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13th International Conference of the European Society for Precision Engineering & Nanotechnology

Monday 27th May to Friday 31st May 2013 Berlin, Germany

conferenceproceedings

Volume 1

www.berlin2013.euspen.eu

Sponsored by:

2013

Conference_Proceedings_vol_1.indd 1 10/05/2013 1237pm

Proceedings of the 13th international conference of the european

society for precision engineering and nanotechnology

May 27

th – 31

st 2013

Berlin, Germany

Volume I Editors: R. Leach P. Shore Proceedings Compilation: T. Horwood D. Nyman D. Phillips N. Williams

Proceedings of the euspen 13th International Conference

Volume l - Volume ll

Reviewed by: Mr. D. Arneson Prof. R. W-B. Lee Mr. S. Azcarate Dr. T. Lundholm Dr. ir. D. Brouwer Dr. S. Ludwick Dr. K. Beckstette Mr. P. Martin Prof. L. Blunt Prof. L. Mattsson Dr. H. Bosse Prof. G. McFarland Prof. C. Brecher Prof. P. McKeown Prof. W. Brenner Dr. K. Monkkonen Prof. E. Brinksmeier Mr. P. Morantz Prof. S. Büttgenbach Prof. T. Moriwaki Mr. K. Carlisle Prof. R. Munnig Schmidt Dr. S. Carmignato Dr. W. Preuss Dr. K. Carneiro Dr. A. Rankers Prof. K. Cheng Prof. D. Reynaerts Prof. D. Chetwynd Dr. O. Riemer Dr. P. Comley Dr. J. Roblee Prof. J. Corbett Prof. R. Schmitt Prof. G. Davies Dr. Ing. H. Schwenke Prof. L. De Chiffre Prof. P. Shore Dr. P. de Groot Prof. A. Slocum Dr. C. During Prof. ir. H. Soemers Mr. P. Eklund Dr. H. Spaan Dr. W. T. Estler Dr. S. Spiewak Dr. C. Evans Dr. P. Subrahmanyan Dr O. Falkenstörfer Prof. K. Takamasu Dr. G. Florussen Prof. Y. Takeuchi Prof. A. Forbes Dr. G. Tosello Dr. H. Haitjema Mr. M. Tricard Prof. H. Hansen Prof. E. Uhlmann Dr. S. Henein Prof. H. Van Brussel Prof. R. Hocken Prof. J. van Eijk Dr. A. Hof Dr. M. Verdi Dr. W. Holzapfel Dr. D. Walker Dr. A. Islam Dr. C. Wenzel Prof. S. W. Kim Dr J. Yagüe-Fabra Dr. W. Knapp Prof. K. Yamamura Dr. L. Kudla Mr. M. Zatarain Prof. R. Leach Prof. S. Zelenika

Published by euspen ISBN 13: 978-0-9566790-2-4

Printed in Netherlands May 2013 © euspen Headquarters Sieca Repro Turbineweg 20

Building 30, Cranfield University, Bedford, MK43 0AL

2627 BP, Delft Tel: 0044 (0) 1234 754154 Website: www.euspen.eu

Information correct at time of printing and may be subject to change

http://www.euspen.eu/

Foreword We are delighted to welcome our delegates to Berlin for euspen’s 13

th

International Conference and Exhibition. Germany’s capital city is one of the

most vibrant cities in the world. Undoubtedly it is a city rich in culture,

politics, media and science.

Berlin is a place of advancement in all respects, home to leading

universities, research institutes and many of the world’s leading industrial

organisations. In regard to euspen’s field of relevance, Berlin is hugely

significant, being the home to many eminent scientists and leading

advanced engineering companies. Names such as Einstein, Planck, Prandtl,

Siemens and HEIDENHAIN give testament to such a suggestion.

In the midst of a period of mixed economic fortune, it is interesting to

recognise Germany’s resilience to the present financial down turn alongside

the same resilience seen by many of the high technology companies which

form euspen’s own industrial community. A common theme is perhaps the

priority placed on high value manufacturing capability applied to leading

edge products and services. Alongside great concern of economic situation,

recent years have seen raised concern for sustainable and safe energy

generation. It is clear for all that Germany has taken strong positions in this

area. So for Berlin we are especially pleased to have an Energy Focused

Keynote and special Renewable Energy session.

Over 160 papers have been selected by the International Scientific

Committee for presentation in Berlin. These papers introduce new key

enabling technologies and ideas relevant for high value manufacturing and

nanotechnology dependent product development. With more than 40 leading

organisations presenting their latest innovations and products the exhibition

will bring leading researchers and industrialists together to nurture new ideas

and opportunities.

Here in Berlin our headline sponsor is HEIDENHAIN. We would like to

express our gratitude for their support of this event and for their long term

support of student scholarships. We would also like to thank our other

sponsors: ASML for the welcome reception, Olympus in regards to the

3

conference proceedings and Physik Instrumente GmbH for the conference

bags.

We thank those involved in scientific reviewing for their exacting work and

congratulate those successful through this review process. Finally, we

acknowledge and deeply appreciate the work of the Session Chairmen for

their help and assistance in defining the presentation programme.

euspen, together with our local hosts, the Fraunhofer Institute for Production

Systems and Design Technology, will ensure you enjoy the vibrant

ambience of Berlin and become acquainted with some local specialty cuisine

and history.

We very much look forward to meeting with you here in Berlin.

Berlin, May 2013

Paul Shore

euspen President

on behalf of euspen Council

4

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5

Contents

Volume I Keynote: Keynote

19

Session 0: Advanced Optics Technology

25

Session 1: Precision Engineering Advancements Enabling Progress in Energy Technologies

51

Session 2: Nano & Micro Metrology

69

Session 3: Ultra Precision Machines & Control

165

Session 4: High Precision Mechatronics

273

7

Proceedings of the 13th euspen International Conference – Berlin – May 2013

K3 Keynote 3: “Industrial Measurement – A Reflection on Progress from 1960 to 2060” Mr Nick Orchard, Rolls Royce, UK

21V1

Oral Session 0: Advanced Optics Technology

O0.1 State-Of-The-Art X-Ray Optical Systems and their Fabrication A. Erko Institute for Nanometre Optics and Technology, Helmholtz Zentrum Berlin,Germany

27V1

O0.2 Optical Glass Grinding with Laser Structured Coarse-Grained Diamond Wheels B. Guo, Q. L. Zhao, W. Zhang, CPE-Center for Precision Engineering, School of Mechatronics Engineering, Harbin Institute of Technology,China

31V1

O0.3

Manufacturing of Freeform Mirror by Milling and Altering its Optical Characteristics by ALD SiO2 Coating J. Mutanen

1 , J. Väyrynen

2 , S. Kivi

3 , M. Toiviainen

3 ,

J. Laukkanen1 , P. Pääkkönen

1 ,T. Itkonen

1 , A. Partanen

1 ,

M. Juuti3 , M. Kuittinen

1 , K. Mönkkönen

2

1University of Eastern Finland, Joensuu, Finland

2Karelia University of Applied Sciences, Joensuu, Finland

3VTT Technical Research Centre of Finland, Kuopio, Finland

35V1

O0.4 Photonic Flip-Flop Based Solutions to Overcome Memory-Wall Challenges P. Tcheg

1, B. Wang

1, M. Palandöken

1,T. Tekin

1,2

1Forschungsschwerpunkt Technologien der Mikroperipherik, TU-

Berlin 2Fraunhofer-Institut für Zuverlässigkeit und Mikrointegration (IZM)

39V1

Oral Session 1: Precision Engineering Advancements Enabling Progress in Energy Technologies

O1.1 Precision Engineering for Concentrating Solar Power (CSP) Applications C Sansom

1, P Comley

1 , P King

1 , N Macerol

1

1Cranfield University, UK

53V1

O1.2 Photocatalytic Activity Influenced by Thickness of TiO2 Measured in Nano and Macro Scale S. Daviðsdóttir

1 K. Dirscherl

2 R. Shabadi S. Canulescu

1

R. Ambata1

1DTU and Denmark

2DFM and Denmark

57V1

8

http://en.hit.edu.cn/sme/



O1.3 An Unconventional Experimental Setup for Testing Cutting Performance/ wear Resistance of Diamond Cutting Wires V. Herold

1 S. König

1 M. Berg

2

1University Jena, Institute for Materials Science and Technology,

Germany 2j-fiber GmbH, Germany

61V1

Oral Session 2: Nano and Micro Metrology

O2.1 In-line Metrology of Functional Surfaces with a Focus on Defect Assessment on Large Area Roll to Roll Substrates L. Blunt

1 , L Fleming

1 , M. Elrawemi

1 , D. Robbins

2 ,

H. Muhamedsalih1

1University of Huddersfield, UK,

2 Centre for Process Innovation, Sedgefield, UK

71V1

O2.2 High-resolution Investigation and Application of Diamond Coated Probing Spheres for CMM- and Form Metrology M. Neugebauer, S. Bütefisch , T. Dziomba , S. Koslowski , H. Reimann Physikalisch-Technische Bundesanstalt (PTB), Germany

75V1

O2.3 Validation of On-machine Microfeatures Volume Measurement Using Micro EDM Milling Tool Electrode as Touch Probe G. Tristo

1, M. Balcon

1, S. Carmignato

2, G. Bissacco

3

1Department of Industrial Engineering, University of Padua, Italy

2Department of Management and Engineering, University of Padua,

Italy 3Department of Mechanical Engineering, Technical University of

Denmark, Denmark

79V1

O2.4

Virtual CMM Method Applied to Aspherical Lens Parameters Calibration A. Küng, A. Nicolet, F. Meli Federal Institute of Metrology METAS

83V1

Oral Session 3: Ultra Precision Machines and Control

O3.1 Concept for a Miniaturized Machine-Tool-Module for the Manufacturing of Micro-Components Operated at its Resonance Frequency C. Oberländer

1, J.P. Wulfsberg

1

1Helmut-Schmidt-University, University of the Federal Armed Forces

Hamburg, Germany

167V1

O3.2 Concrete Based Parts for High Precision Applications C. Hahm

1, R. Theska

1, K. John

1, A. Flohr

2, A. Dimmig-Osburg

2

1Technische Universität Ilmenau, Germany

2Bauhaus-Universität Weimar, Germany

171V1

9


O3.3 Fast Nanometer Positioning System by Combining Fast Resonant Mode and Accurate Piezostack Direct Drive A. Santoso, J. Peirs, F. Al-Bender, D. Reynaerts KU Leuven, Department of Mechanical Engineering, Belgium

175V1

O3.4 Towards the Realization of the New INRIM Angle Comparator M. Pisani and M. Astrua Istituto Nazionale di Ricerca Metrologica, INRIM, Italy

179V1

O3.5 Geometrical-based approach for flexure mechanism design T.J. Teo

1 , G. Z. Lum

1,2,3 , G.L. Yang

1 , S. H. Yeo

2 , M. Sitti

3

1Singapore Institute of Manufacturing Technology, Singapore

2Nanyang Technological University, Singapore

3Carnegie Mellon University, United States.

184V1

Oral Session 4: High Precision Mechatronics

O4.1 FEM Model Based POD Reduction to Obtain Optimal Sensor Locations for Thermo-elastic Error Compensation J. van der Sanden, P. Philips Philips Innovation Services, The Netherlands

275V1

O4.2 2-DoF Magnetic Actuator for a 6-DoF Stage with Long-stroke Gravity Compensation R. Deng, J. W. Spronck, A. Tejada, R. H. Munnig Schmidt PME: Mechatronic System Design, Delft University of Technology, The Netherlands

279V1

O4.3 Highly Accurate Passive Actuation System S. A. J. Hol

1, J. Huang

1, W. Zhou

1, M. Koot

1, H. Vermeulen

1,

J. van Eijk2, R. Munnig-Schmidt

3

1ASML BV, The Netherlands

2MICE BV, The Netherlands

3Delft University of Technology, Mechatronic Systems Design, The

Netherlands

283V1

O4.4 Design and Fabrication of a Novel Centimeter Scale Three Dimensional Silicon Tip, Tilt and Piston Mirror Mechanism J. Kruis

1,2, F. Barrot

1, L. Giriens

1, D. Bayat

1, R. Fournier

1,

S. Henein2, S. Jeanneret

1

1Centre Suisse d’Electronique et de Microtechnique (CSEM),

Switzerland 2École Polytechnique Fédérale de Lausanne (EPFL), Switzerland

288V1

O4.5 Superstructures control with active tie rods C. Collette, D. Tshilumba, L. Fueyo-Rosa University of Brussels, Belgium

292V1

Posters Sessions 0 - 4 Session 0: Advanced Optics Technology

10


P0.01

High Precision Injection Moulding of Freeform Optics with 3D Error Compensation Strategy L. Dick

1,2, S. Risse

3, A. Tünnermann

2,3

1JENOPTIK Polymer Systems GmbH, Germany

2Friedrich Schiller University Jena, Abbe Center of Photonics,

Institute of Applied Physics, Germany 3Fraunhofer Institute for Applied Optics and Precision Engineering

IOF, Germany

43V1

P0.02

Integration Platform of Dual Wavelength Signal Source for 120GHz Wireless Communication Systems M. Palandöken

1, T. Tekin

1, 2

1Forschungsschwerpunkt Technologien der Mikroperipherik, TU-

Berlin 2Fraunhofer-Institut für Zuverlässigkeit und Mikrointegration (IZM, ,

Germany)

47V1

Session 1: Precision Engineering Advancements Enabling Progress in Energy Technologies

P1.01 Fabrication of Freeformed Blazed Gratings by Ultraprescision Machining K. Haskic

1*, S. Kühne

1*, S. Lemke

2, M. Schmidt

1

1Technische Universität Berlin, Institut Für Werkzeugmaschinen und

Fabrikbetrieb (IWF), Fachgebiet Mikro- und Feingeräte (MFG), Germany 2Helmholtz-Zentrum Berlin für Materialien und Energie (HZB),

Institut Nanometeroptik und Technologie (G-INT), Germany *Equally contributing

65V1

Session 2: Nano and Micro Metrology

P2.01 3D Shape Measurement Under Multiple Refraction Condition Using Optical Projection Method Y. Uchida, R. Kamei, Y. Higashio Department of Mechanical Engineering, Aichi Institute of Technology,Japan

87V1

P2.02 Elastic Behaviour of Millimetre-scale Polymeric Triskelion-like Flexures D.G. Chetywnd, Z. Davletzhanova, Y. Kogoshi, H. ur Rashid School of Engineering, University of Warwick, UK

91V1

P2.03 Scanning Results and Repeatability Testing of the TriNano Ultra Precision CMM A.J.M. Moers

1, M.C.J.M. van Riel

1,2, E.J.C. Bos

1

1Xpress Precision Engineering, The Netherlands

2Eindhoven University of Technology, The Netherlands

95V1

P2.04 Distance Ranging Using Original Fiber-optic Interferometer J K. Thurner, P.-F. Braun, K. Karrai attocube systems AG, Königinstrasse München, Germany

99V1

11


P2.05 Design of a Nanometer-accurate Air Bearing Rotary Stage for the Next Generation Nano-CT Scanners S. Cappa, D. Reynaerts, F. Al-Bender KU Leuven, Department of Mechanical Engineering, Belgium

103V1

P2.06 Practical Method for Determining the Metrological Structure Resolution of Dimensional CT S. Carmignato

1, P. Rampazzo

1, M. Balcon

2, M. Parisatto

3

1University of Padova, Department of Management and

Engineering, Italy 2 University of Padova, Department of Industrial Engineering, Italy

3 University of Padova, Department of Geosciences, Italy

107V1

P2.07 Traceable Profilometer with a Piezoresistive Cantilever for

High-aspect-ratio Microstructure Metrology

M. Xu, U. Brand, J. Kirchhoff Physikalisch-Technische Bundesanstalt (PTB),Germany

111V1

P2.08 Verification of Thickness and Surface Roughness of a Thin Film Transparent Coating K. Mohaghegh

1, H.N. Hansen

1, H. Pranov

2, G. Kofod

2

1Technical University of Denmark, Denmark

2InMold Biosystems, Denmark

115V1

P2.09 Measurement and Evaluation Processes for Inner Micro Structures T. Krah

1, A. Wedmann

1, K. Kniel

1, F. Härtig

1

1Physikalisch-Technische Bundesanstalt, Braunschweig und Berlin,

Germany

120V1

P2.10 Quantitative Assessment of Nano Wear of DLC Coated Samples using AFM and Optical Confocal Microscopy G. Dai

1, F. Pohlenz

1, H. Bosse

1, A. Kovalev

2, D. Spaltmann

2, M.

Woydt2

1Physikalisch-Technische Bundesanstalt (PTB), Braunschweig,

Germany 2 Federal Institute for Materials Research and Testing (BAM), Berlin,

Germany

124V1

P2.11 Measurement Setup for the Experimental Lifetime Evaluation of Micro Gears G. Lanza

1, B. Haefner

1

1wbk Institute of Production Science, Karlsruhe Institute of

Technology (KIT), Germany

128V1

P2.12 3D-Reconstruction of Microstructures on Cylinder Liners F. Engelke, M. Kästner, E. Reithmeier Institute of Measurement and Automatic Control – Leibniz Universität Hannover

132V1

12


P2.13 A Self-calibration Method for the Error Mapping of a 2D Precision Sensor M. Valenzuela

1, M. Torralba

2, J.A. Albajez

1, J.A. Yagüe

1, J.J.

Aguilar1

1I3A, University of Zaragoza, Spain

2Centro Universitario de la Defensa, Zaragoza, Spain

136V1

P2.14 Reaming in Microscale of Titanium and Titanium Alloys D. Biermann, J. Schlenker Department of Machining Technology, Technische Universität Dortmund, Germany

140V1

P2.15 Investigation of Stylus Tip-size Effects in Surface Contact Profilometry K. T. Althagafy¹’², D G Chetwynd¹ ¹School of Engineering, University of Warwick, Coventry, UK ²Umm AlQura University, Saudi Arabia

144V1

P2.16 ISO Compliant Reference Artefacts for the Verification of Focus Varation-based Optical Micro-co-ordinate Measuring Machines F. Hiersemenzel

1, J. D. Claverley

2; J. Singh

1, J. N. Petzing

1, F.

Helmli3, R. K. Leach

2

1Loughborough University, Loughborough, UK;

2National Physical Laboratory, Teddington, UK;

3Alicona Imaging GmbH, Graz, Austria

148V1

P2.17 Acoustic Emission-based Micro Milling Tool Contact Detection as an Integrated Machine Tool Function E. Uhlmann, N. Raue, C. Gabriel Department of Machine Tools and Factory Management, Chair for Manufacturing Technology, Technische Universität Berlin, Germany

152V1

P2.19 Dimensional verification of high aspect ratio micro structures using FIB-SEM Y. Zhang

1, H.N. Hansen

1

1 Department of Mechanical Engineering, Technical University of

Denmark, Denmark (DTU)

156V1

P2.20 Setting-up Kriging-based Adaptive Sampling in Metrology D. Romano

1, R. Ascione

2

1University of Cagliari, Italy

2ENEA, Italy

160V1

Session 3: Ultra Precision Machines and Control

P3.01 A New Approach on Reducing Thermal Impacts on High Precision Machine Tools M. Fritz

1, Dr. D. Janitza

1

1KERN Microtechnik GmbH, Germany

188V1

13


P3.02 Long Range Precision Stage Using Multi Bar Mirrors S. Woo, D. Ahn, J. Park, D. Gweon Korea Advanced Institute of Science and Technology (KAIST), Republic of Korea

192V1

P3.03 Feasibility Study on a Spindle Supported by High Stiffness Water Hydrostatic Bearings for Ultra-precision Machine Tool Y. Nakao

1, K. Yamada

1, K. Wakabayashi

1, K. Suzuki

1

1 Kanagawa University, Japan

196V1

P3.04 Investigations of a Small Machine Tool with CFRP-frame 1H.-W.

Hoffmeister,

1A. Gerdes,

2A.Verl,

2K.-H. Wurst,

2T. Heinze,

2C. Batke

1TU Braunschweig, Institute of Machine Tools and Production

Technology, Braunschweig, Germany 2Universität Stuttgart, Institute for Control Engineering of Machine

Tools and Manufacturing Units, Stuttgart, Germany

200V1

P3.05 The Dynamic Design of an Ultra-precision Machine Tool Used for Larger KDP Crystal Machining Y. Liang, W. Chen*, Y. Sun, Q. Zhang, F. Zhang Center for Precision Engineering, Harbin Institute of Technology, Harbin, China

204V1

P3.06 Investigation of Micro-optic Polishing Characteristics by Vibration-assisted Polishing J. Guo

1*, Y. Yamagata

1, H. Suzuki

1, 2, S. Morita

1,T. Higuchi

3

1The Institute of Physical and Chemical Research (RIKEN), Wako,

Saitama, Japan 2Department of Mechanical Engineering, Chubu University,

Kasugai, Aichi, Japan 3Department of Precision Engineering, The University of Tokyo,

Tokyo, Japan

208V1

P3.07 Parameter Determination for an Electromechanical Model of a Displacement-Amplified Piezoelectric Actuator J.H. Liu

1, W. O’Connor

1, E. Ahearne

1 and G. Byrne

1

1 School of Mechanical and Materials Engineering, University

College Dublin,Ireland

212V1

P3.08 Ultraprecise Positioning Mechanism with 3-DOF Over a One-millimeter Stroke Using Monolithic Flexure Guide and Electromagnetic Actuator S. Fukada

1, T. Matsuda, Y. Aoyama, T. Kirihara

1Shinshu University, Japan

216V1

P3.09 Design and construction of a novel assisted tool-holder L. Javarez Jr

1, J.G. Duduch

1, R.G. Jasinevicius

1, A.M. Gonçalves

1

1University of São Paulo, Brazil

220V1

14


P3.11 Development of a Vertical-spindle Rotary Surface Grinding Machine for Large Scale Silicon-wafers – Machine Specifications and Performance of Rotary Work Table A.Yui

1, A.Honda

1, S.Okuyama

1, T.Kitajima

1, G.Okahata

1, H.Saito

2,

A.H.Slocum3

1National Defense Academy, Japan

2Okamoto Machine Tool, Japan

3Massachusette Institute of Technology, USA

224V1

P3.12 Band-limited Cutting Force Control in Ultra-precision Turning K. Enomoto

1, Y. Kakinuma

1

1Department of System Design Engineering, Keio University, Japan

228V1

P3.13 Ultra Precision Process Monitoring C. Brecher

1, D. Lindemann

1, A. Merz

1, C. Wenzel

1

1Fraunhofer Institute for Production Technology IPT Germany

232V1

P3.14 Analysis of Mutual Influences of Control, Feedback and Servo Drive Systems for Ultra Precision Machining C. Brecher

1, D. Lindemann

1, C. Wenzel

1

1Fraunhofer Institute for Production Technology IPT, Germany

236V1

P3.15 Determining the Random Measurement Errors of a Novel Moving-scale Measurement System with Nanometre Uncertainty J. N.Bosmans

1, J. Qian

1, D. Reynaerts

1

1KU Leuven, Department of Mechanical Engineering, Belgium

240V1

P3.16 An Approach to the Optimal Observer Design with Selectable Bandwidth I. Furlan, M. Bianchi, M. Caminiti, G. Montù University of Applied Sciences of Southern Switzerland, Manno, Switzerland

244V1

P3.17 Bandwidth Increase for Plate-like Structures by Adding Mechanical Dampers C.A.M. Verbaan

1, P.C.J.N. Rosielle

1, M. Steinbuch

1

1 Control Systems Technology group, Department of Mechanical

Engineering, Eindhoven University of Technology, The Netherlands

248V1

P3.18 A Parallelism Alignment Mechanism for Nanoimprint Lithograph with Large Imprinting Force W.J. Chen, W. Lin, G.L. Yang Singapore Institute of Manufacturing Technology (SIMTech), Singapore

252V1

P3.19 Design and Performance of a 6 DOF Hybrid Hexapod N.L. Brown

1, C.W. Hennessey

1

1ALIO Industries, USA

256V1

15


P3.21 Concept Design of a 5-axis Portable Milling Machine for the In-situ Processing of Large Pieces J. Eguia

1, O. Gonzalo

1, M. San Martín

1, S. Ilhenfeldt

2

1IK4 - TEKNIKER, Spain

2Fraunhofer IWU – Germany

260V1

P3.23 Using Boron Doped Diamond Foils for Fabrication of Micro Cavities with EDM K E. Uhlmann

1, M. Langmack

1, J. Fecher

2, S. M. Rosiwal

2,

R. F. Singer2

1Institute for Machine Tools and Factory Management,

Technische Universität Berlin, Germany 2Institute of Science and Technology of Metals (WTM), University of

Erlangen-Nuremberg, Erlangen, Germany

264V1

P3.24 Design and Optimization of Flexure-Based Micro-manipulator for Optics Alignment C. Brecher, N. Pyschny, T. Bastuck Fraunhofer Institute for Production Technology IPT, Germany

268V1


P4.01 Modelling Lateral Web Dynamics for R2R Equipment Design B. J. de Kruif, H. E. Schouten TNO, The Netherlands

296V1

P4.02 Design of an Active Magnetic Stabilizer of the Dynamic Behaviour of High Speed Rotors E. Brusa Dept. Mechanical and Aerospace Engineering, Politecnico di Torino, Italy

300V1

P4.03 Physical and Phenomenological Simulation Models for the Thermal Compensation of Rotary Axes of Machine Tools M. Gebhardt, S. Capparelli, M. Ess, W. Knapp, K. Wegener Institute of Machine Tools and Manufacturing (IWF), ETH Zurich, Switzerland

304V1

P4.04 Compact Translatory Actuator with Moving Magnets and Flexure Guide for Versatile Applications T. Bödrich, F. Ehle, J. Lienig Technische Universität Dresden, Institute of Electromechanical and Electronic Design, Germany

310V1

P4.05 Displacement of a 6-DOF Inchworm-based Parallel Kinematic Stage A. Torii, R. Kamiya, K. Doki, A. Ueda Dept. of Electrical and Electronics Eng., Aichi Institute of Technology, Japan

314V1

16


P4.07 Increased Productivity due to Jerk-decoupled Feed Axes of the 5-Axes Milling Machine “Neximo” B. Denkena, K. Litwinski, O. Gümmer Institute of Production Engineering and Machine Tools (IFW), Leibniz Universitaet Hannover, Germany

318V1

P4.08 Design and Optimization of a 3-DOF Planar MEMS Stage with Integrated Thermal Position Sensors B. Krijnen

1,2, K. R. Swinkels

1,2, D. M. Brouwer

1,2, J. L. Herder

2

1DEMCON Advanced Mechatronics, The Netherlands

2Mechanical Automation & Mechatronics, University of Twente,

The Netherlands

322V1

P4.10 Sensorless Monitoring of Machining Torque on Tilting Platform Driven by Hybrid Actuator H. Yoshioka

1, M. Hayashi

2, H. Sawano

1, H. Shinno

1

1Tokyo Institute of Technology, Japan

2The University of Tokyo, Japan

326V1

P4.11 Self-tuning Dynamic Vibration Absorber for Machine Tool Chatter Suppression G. Aguirre

1, M. Gorostiaga

1, T. Porchez

2, J. Muñoa

1

1IK4-IDEKO, Spain

2CEDRAT TECHNOLOGIES, France

330V1

P4.12 Design and Control of a Through Wall 450 mm Vacuum Compatible Wafer Stage D. Laro

1, E. Boots

2, J. van Eijk

2,3, L. Sanders

1

MI-Partners, The Netherlands1

TU Delft, The Netherlands2

MICE BV, The Netherlands3

334V1

P4.13 Driving a Femtosecond Machined Tactile Scanning Probe Stage in the 100 µm Range D. F. Vles, F. G. A. Homburg Eindhoven University of Technology, The Netherlands

338V1

17

Keynote

19


Industrial measurement – a reflection on progress from

1960 to 2060

N B Orchard

Rolls-Royce plc, UK

[email protected]

Abstract

It is just over 50 years since the first commercial laser was produced, and since then

the laser has been an increasingly important tool for the metrologist. As well as the

laser a number of other inventions have contributed to some major advances in the

technology that we use to measure the products that we make. We have moved from a

predominantly manual process using mechanical gauges and reference artefacts, to a

process that is largely automated, using computers to control coordinate measuring

machines, or systems that gather vast numbers of coordinates using light or X-rays to

digitise the objects of scrutiny. However, as with many things, quantity does not

necessarily imply quality, and it would be easy to be misled into thinking that our

ability to measure things in the 21st century is way in advance of our technology of 50

years ago. This talk will discuss aspects of what we call „progress‟ and how we

measure it, and look forward to what developments we might see or want in the next

50 years.

1 The early years of accuracy

There are two key requirements for any system of measurement: the first is a baseline

standard for the unit of measurement that can be accurately reproduced, and the

second is a method of using that standard to assess the measurand of interest. In order

to put things into perspective it would perhaps be worth going back a little further in

order to judge the progression of measurement technology. If I may be permitted to

skip the first several thousand years of measurement technology, the early

development of what we would recognise today as precision metrology probably

started in the 18th Century. The demand for accurate measurement is driven by the

things that we want to make. Until the 18th Century the things that people wanted to

make could on the whole be made without accurate measurement, things like houses,

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furniture, carriages and wagons. It was the advent of „machines‟ that prompted the

measurement developments, machines like clocks and watches, telescopes and

surveying equipment, cotton and wool spinning and weaving machinery, and then of

course the development of guns, engines and vehicles. Instrument makers such as

John Bird and Edward Troughton were able to make replica yard scales in the mid to

late 1700‟s that varied by around 0.003”. By the early 1800‟s calipers were available

that claimed to be accurate to 0.001”, and micrometer scales that could be read to

0.0001”, although the absolute accuracy of these could be debated.

2 Interferometry

Although the next 150 years saw continued growth in the range of measurement

devices, the ability to measure real objects with greater accuracy did not change as

much as one might have expected. Michelson had lead the way in suggesting a new

fundamental length standard based on the speed of light, and demonstrated

interferometry to measure the metre in 1893, but it wasn‟t until the invention of the

HeNe laser in Bell Laboratories in the early 1960‟s that a practical coherent light

source became available. Airborne Instruments Labs produced the first commercial

laser based displacement interferometer in 1964, followed by Perkin-Elmer‟s

“Lasergage” homodyne interferometer in 1968, and the HP552A laser interferometer

in 1970. It was also only in 1960 that the physical artefact definition of the metre was

replaced by a definition based on the wavelength of a stable light source. While these

laser systems have led to major steps forward in distance measurement, and allowed

the re-definition of the metre based on the second and the speed of light, they in

themselves do not enable us to measure the real objects that we make.

3 Coordinate Measurement Machines

It is perhaps coincidence that the second most influential invention relating to

dimensional measurement was also produced at the same time as the laser. The co-

ordinate measuring machine was introduced by DEA in Italy and by Ferranti in the

UK in 1959/1960, albeit with some debate about who was first. The early machines

were entirely manually operated but had digital readouts of the solid probe position.

Developments of the technology over the last 50 years have produced the machines

that we have today, with fully automatic computer control, high-speed scanning

22


probes and the ability to be programmed from CAD. There is now very little that we

make that cannot be measured on a CMM.

4 Non-contact measurement

In tandem with the development of the contact probe CMM has been the development

of non-contact techniques such as photogrammetry, laser triangulation and structured

light systems. Photogrammetry has been used virtually since the invention of

photography in the mid 19th Century, and has been used extensively for mapping,

surveillance and measurement of (generally) large objects. However, it is only with

the advent of digital cameras and high-speed computing that it has been possible to

automate the use of photogrammetry, and also to bring it into the realm of meso-scale

metrology. Likewise, structured light systems have only been feasible since the

invention of the digital camera. Non-contact camera-based measurement systems

always have a trade-off between field of view and resolution/accuracy, such that they

have never really been competitive with CMMs for absolute measurement accuracy.

Their main strength has been in the quantity of data that they can provide, so where a

CMM may be able to measure say 50 points in a minute, a structured light system can

measure 5,000,000 points. The fact that the non-contact system can give the viewer a

full image of the object‟s surface and its deviations may be more valuable than the

more accurate, but limited data output of the CMM. While the standard shop CMM is

probably nearing the limit of its speed and accuracy potential, the constant

development of cameras and computers means that the structured light systems still

have significant development potential.

5 The future

So what developments can we expect in metrology in the next 50 years? Apart from

the non-contact development potential just mentioned, it is likely that the most

opportunities lie in some of the ways that we use metrology and in our knowledge of

the process. At the moment we tend to see product measurement as a necessary

overhead, and there is a strong desire to minimise the amount of measurement that we

do, but without compromising product quality. What can we do to improve our

machining processes, or our knowledge of them, so that we can be confident that the

process is producing conforming output without the need to measure everything?

23


Can we improve our understanding of the behaviour of both our machines and our

measurement systems so that we could be much better at predicting what can or

cannot be made capably on a particular machine? Can we build all our best practice

experience into the programming systems that we use to drive the machining and

measurement systems?

One thing is for certain, and that is that our shop floors will look very different in

2060 from the way they look today. Exactly what the changes will be I can‟t predict

very well, but I can predict that it should give a lot of people a lot of fun getting there.

24

Session 0: Advanced Optics Technology

25


State-Of-The-Art X-Ray Optical Systems and their

Fabrication

A. Erko

Institute for Nanometre Optics and Technology, Helmholtz Zentrum Berlin, Albert-

Einstein-Str. 15, 12489 Berlin, Germany

[email protected]

Abstract

Two examples of modern developments in the field of X-ray optical elements and

systems in Berlin-Adlershof are presented: focusing optical systems on the basis of

glass capillaries, developed in the IfG Institute for Scientific Instrumentation GmbH,

and diffractive optical systems on the basis of diffraction gratings and reflection zone

plates, developed in the Institute for Nanometer Optics and Technology of the

Helmholtz Zentrum Berlin.

1 Specialized Glass Capillary Optics

Glass capillary optics can be used for numerous different X-ray analytical methods

such as XRF or XRD, µEXAFS, µXANES and since very recently also full field X-

ray fluorescence in energy interval of 1-30 keV. Improvements of the production

technology for mono-capillary optics made it also possible to reach routinely spot

sizes down to 1 µm when used with synchrotron radiation (SR) source and down to

10 µm with laboratory sources. These lenses are used for X-ray spectroscopy with

synchrotron radiation at the BESSY II facility as well as in laboratory-scale

instrumentation, for example XRF applications in scanning electron microscopes.

For applications in micro X-ray fluorescence analysis, a new generation of poly-

capillary optics was produced with improved physical parameters such as spot sizes

on the order of 10 µm for MoK and intensity gains of more than 10 000. New

types of optics are required for the full-field colour X-ray camera was recently

developed. This large exchangeable poly-capillary array (12x12 mm) in front of the

inlet aperture of the camera conducts X-ray photons from the probe to the energy

dispersive pixels on a pn-CCD. This transmission of X-ray photons inside the

capillary channels enables correct imaging of the sample on the pixels without any

27



cross views. It would make real-time visualization of the element distribution in a

sample possible without scanning system.

Figure 1. Glass capillary X-Ray optics produced by IfG Institute for Scientific

Instrumentation GmbH. (A. Bjeoumikhov)

The aim of future developments in capillary optics is to further decrease focal spot

sizes while increasing brilliance. Such parameters can only be realised if

corresponding high brilliant micro-focus sources are available and a high quality of

capillary optics can be guaranteed. New developments for capillary optics were

carried out by changing the capillary diameters as well as using new glass types to

improve the transmission qualities. It was shown that poly-capillary optics can be

flexibly adapted to concrete applications.

2 Advanced Diffractive Optical Systems

The Institute for Nanometer Optics and Technology (INT) has extensive experience in

micro fabrication (technology group) and X-ray beamline optics design (optics

group). In spring 2010, the HZB in cooperation with partners: DIOS GmbH and Chair

micro and precision devices - TU Berlin decided to build up its own technology centre

for diffractive x-ray optics design and fabrication. Besides traditional lamellar

diffraction gratings for synchrotron radiation applications, we have developed

technology for advanced x-ray optical components such as lamellar and blazed

gratings as well as reflection zone plates (RZP) for monochromatization and

spectroscopy in the VUX energy range.

28


2.1 Fabrication technologies

The main technology for grating fabrication at the HZB is mechanical ruling. The

workhorse of this technology is the old C. Zeiss ruling engine, GTM-6. After its

transportation to Berlin, the machine had to be repaired. Mechanical and electronic

components were maintained and, if necessary, replaced. The first ruled gratings

were produced in December 2011. In the meantime, the engine has been installed in

a thermo-stabilized cleanroom environment. The GTM-6 is able to process

substrates up to 170 mm length. In Figure 2 is shown a grating produced in March

2013 at the GTM-6. In the following time the ruling process was optimized and

several blazed gratings on silicon substrates were generated with line densities of

650 to 2000 lines/mm. A typical result is shown in figure 2.

A new GTM-24, which will be able to process substrates up to 600 mm in length, is

currently under construction. It will be delivered in summer 2013.

Figure 2. Diffraction grating, L: 96 mm, W: 16 mm,

600 lines/mm, Blaze: 5.4° 57h ruling duration. 57600 lines (T. Zeschke)

The wet chemical etching of asymmetrically cut mono-crystalline silicon is another

effective fabrication method for blazed gratings. The patterning of the necessary

etching mask can be done by e-beam writing or holography. Precondition for that is a

precisely cut Si substrate with a super polished surface. Accurate surface cleaning is

necessary for the KOH etching process. A major challenge for the grating fabrication

by wet chemical etching is the adjustment of the etching mask to the crystal planes.

The Si etching method is a very cost-efficient process for generating high quality

blazed gratings.

29


The method of laser holography is established for fabricating all kinds of laminar

gratings by lithography. Resist patterning at the HZB is performed by laser

interference lithography using an UV laser with 442 nm wavelength. When applying

the laser optics we can presently expose circular areas of about 100 mm. However,

we intent to overcome this present limit by transferring our set-up to a larger optical

table.

The structure of reflection zone plates (RZP), shown in figure 3, was made by using

high-voltage electron beam lithography (VISTEC EBPG 5000plusES) and reactive

ion etching techniques. We used a super-polished substrate, patterned and gold

coated. A RZP with lateral dimensions of 80 mm × 2.4 mm, a lamellar profile of 13

nm and a minimum zone width of 70 nm was produced on the substrate surface.

Figure 3. An SEM image of the spectrometer structure (Si substrate, Au coating).

3. Acknowledgement

This work is funded by the European Community with money from the European

Regional Development Fund (ERDF) under contract No. 20072013 2/43 as well as

the BMBF project 05K12CB4.

We acknowledge contributions by the IfG colleagues A. Bjeoumikhov, S.

Bjeoumikhova N. Langhoff and O. Scharf. As well as IMT HZB colleagues: J.

Buchheim, F. Eggenstein, R. Follath, A. Gaupp, Ph. Goettert, G. Gwalt, K. Haskic,

S. Künstner, St. Kühne, O. Kutz, A. Panner, I. Rudolph, F. Schaefers, T. Selinger, T.

Senn, F. Siewert, D. Stoppel, Ch. Waberski, J. Wolf, T. Wolf, T. Zeschke, I. Zizak.

B. Lochel, A. Firsov, M. Brzhezinskaya, B. Nelles (DIOS GmbH), M. Schmidt †3

(TU Berlin).

30


Optical Glass Grinding with Laser Structured Coarse-

Grained Diamond Wheels

B. Guo, Q. L. Zhao and W. Zhang

CPE-Center for Precision Engineering, School of Mechatronics Engineering,

Harbin Institute of Technology, Harbin, 150001, China

[email protected]

Abstract

Coarse-grained diamond wheels can realize high efficient grinding of optical glass.

However, the subsurface damage will be inevitably introduced by the coarse-grained

wheels. Based on laser machining method, this paper presents a structured coarse-

grained diamond wheel for optical glass grinding. Continuous micro grooves with 10-

15μm width were machined on the peripheral surface of grinding wheel by UV

nanosecond pulsed laser. The protruding part of most diamond grains were cut-

through by grooves. Damage of diamond grains or their falling out was not found

during laser structuring. The results of optical glass grinding tests show than the

subsurface damage depth could be reduced effectual when using the structured

coarse-grained diamond wheels, better surface quality was not however obtained.

1 Introduction

A drawback of using fine-grained diamond wheels to grind optical glass in ductile

mode is the large wheel wear rate caused by the dressing and grinding process, which

limits the achievable figure accuracy and the maximum material removal volume.

Well conditioned coarse-grained diamond wheels featuring grain sizes of approx. 70-

300μm, can be a solution of the wheel wear problem and realize high efficient

precision grinding of hard and brittle materials [1, 2]. However, serious subsurface

damage will be introduced by coarse-grained wheels inevitably [3].

Based on above issue, this paper presents the potential of a structured coarse-

grained diamond wheel with micro grooves on the wheel surface for optical glass

surface grinding aiming to improve the grinding performance, especially reducing

subsurface damage. Firstly, the experimental investigation of coarse-grained diamond

wheel structuring was carried out by UV nanosecond pulsed laser. The morphology

31



of structured wheels was analyzed by SEM. Some structured coarse-grained diamond

wheels with different interval micro grooves were manufactured under the optimal

laser parameters. Then the wheels were conditioned with a metal bond diamond

wheel with ELID method. Finally, the structured coarse-grained diamond wheels

were used in optical glass grinding experiments. Wedge polishing methods were used

to measure subsurface damage depth. The generated surface and subsurface quality of

BK7 samples were characterized by profilometer and SEM. The effect of grinding

wheel grooves interval on achieved surface roughness and subsurface damage depth

were investigated.

2 Structuring of coarse-grained diamond wheel by laser

In this experiment, UV nanosecond pulsed laser are selected as the laser source to

machining micro grooves on coarse-grained diamond wheel. The 1A1 type

electroplated diamond grinding wheels with 150μm grain size were used in this

experiment. A precision spindle was used to rotate the grinding wheel under the laser

source. The micro grooves were parallel with each other. The interval of grooves

could be controlled by a Z slide way. Four structured grinding wheels with different

groove interval were machined by the laser. Half of the cross sectional area of the

grinding wheel was structured with 70μm interval micro grooves. The intervals of

grinding wheel were 30μm, 90μm and 150μm, respectively.

The morphology of structured wheel is shown in Fig. 1. On structured surface, the

continuous grooves were obtained. The width of grooves was 10-15μm. The

protruding part of most diamond grains were cut-through by one or two grooves. The

broken diamond grain and falling off of grain was not found.

Fig. 1 The morphology of structured coarse-grained diamond grinding wheel

With

structuring Without

structuring

32

app:ds:respectively


3 Grinding experiments of optical glass

The experiments of grinding were conducted on a precision grinder MUGK7120X5.

The work piece was optical glass BK7. The grinding parameters were 3000r/min

spindle speed, 2μm depth of grinding and 2mm/min feed rate. Water-base emulsion

was used as a coolant to improve the grinding condition. The angle between the

grinding feed direction and wheel axial direction was 45°.

The ELID assisted conditioning technique with metal-bond diamond truer was

used to condition the structured coarse-grained diamond wheel. A radial run out of

less than 20μm as well as the top-flattened diamond grains of constant wheel

peripheral envelop was generated. The surface morphology of the conditioned

grinding wheel was examined using the SEM, as shown as Fig. 2.

Fig. 2 The morphology of conditioned grinding wheel ×75

The surface roughness Ra values were measured by means of a contact probe

profilometer (Talysurf PGI 1240) on different direction. The results showed the better

surface quality would not be obtained by structured coarse-grained diamond wheel

compare with conditioned coarse-grained diamond wheel. Moreover, the interval of

micro grooves would influence grinding quality. The surface roughness Ra was

improved when the interval decreased.

The subsurface damage was investigated by wedge polishing methods. The

polished surfaces were etched for 3 minutes by NH4HF2 solution. The SEM images

of subsurface damage were shown in Fig. 3. The subsurface damage depth was

improved from 16μm (by original coarse-grained wheel) to 5μm because of the

uniform grain protrusion height due to ELID truing method condition. The subsurface

damage depth induced by the structured wheel was further decreased to less than

3μm, because the coarse grains were “refined” by micro grooves. The subsurface

With

structuring

Without

structuring

33


damage depth seems to reduce with the decreasing interval. At the interval of 30μm,

the subsurface damage depth of 1.5μm was obtained.

(a) Without structuring (b) 150μm interval grooves (c) 30μm interval grooves

Fig. 3 The SEM images of BK7 ground subsurfaces (×2500)

4 Conclusions

The coarse-grained diamond grinding wheels were structured by UV nanosecond

pulsed laser, successfully. Although the better surface quality would not be obtained

by structured coarse-grained diamond wheel compared with conditioned coarse-

grained diamond wheel, the subsurface damage depth could be reduced when using

the structured coarse-grained diamond wheel. The surface roughness and subsurface

damage depth were both reduced with the decreasing interval. For future research, the

structured coarse-grained diamond grinding wheels, which will be conditioned by

ELID before laser machining will be investigated.

References

[1] J.C. Aurich, P. Herzenstiel, H. Sudermann. High-performance dry grinding

using a grinding wheel with a defined grain pattern. CIRP Annals -

Manufacturing Technology, Vol. 57 (2008), p. 357–362.

[2] Q. Zhao, J. Chen, E. Brinksmeier: Precision Grinding of Reaction Bonded

Silicon Carbide Using Coarse Grain Size Diamond Wheels. Chinese Journal of

Mechanical Engineering, Vol. 23 (2010), p. 269-275.

[3] Q. Zhao, E. Brinksmeier, O. Riemer: ELID assisted precision conditioning of

coarse-grained diamond grinding wheel. Key Engineering Materials, Vol. I

(2008), p. 578-583.

34


Manufacturing of Freeform Mirror by Milling and Altering

its Optical Characteristics by ALD SiO2 Coating

J. Mutanen1, J. Väyrynen2, S. Kivi3, M. Toiviainen3, J. Laukkanen1, P. Pääkkönen1,

T. Itkonen1, A. Partanen1, M. Juuti3, M. Kuittinen1, K. Mönkkönen2 1University of Eastern Finland, Department of Physics and Mathematics, Joensuu,

Finland 2Karelia University of Applied Sciences, Joensuu, Finland 3VTT Technical Research Centre of Finland, Kuopio, Finland

[email protected]

Abstract

In this study four aluminium and brass freeform mirrors used as fiber optic

spectrometer probes were micro-milled and ALD SiO2 coating was added. Freeform

surfaces were designed by combining optical modeling with the mechanical structure.

Moore 350 FG ultra precision machine tool was used for milling the freeform parts.

To evaluate the surface roughness of the machining one aluminium and one brass

freeform mirror also contained a 40 mm radius reference lens. The surface

roughness’s of parts containing reference lenses were analysed prior to coating them

with SiO2 on atomic layer deposition (ALD) device. The uncoated and coated

reference lenses were measured with optical profiler. The Ra values of micro-milled

reference lenses on uncoated aluminium and brass surfaces were in good correlation

to theoretical values. On uncoated aluminium mirror surface roughness Ra was 6.8 nm

and for uncoated brass mirror Ra was 5.7 nm. On coated aluminium mirror surface

roughness Ra was 5.2 nm and on coated brass mirror Ra was 5.5 nm. For the testing of

the spectral functionality of the system the mirrors were coupled to a fiber-optic UV-

VIS spectrometer. Spectral measurements were done on coated and uncoated mirrors

with several different coloured reflectance standards. Spectral measurements show

that SiO2 coating affects the reflectance characteristics of both mirror surfaces while

maintaining high reflectivity characteristics.

1 Introduction

The use of state-of-the-art CNC and multi-axis ultra precision diamond machining as

well as optics design and simulation tools has enabled integration of optical functions

directly to the high-accuracy parts of various non-imaging optical systems. The

35



increased accuracy of the manufacturing of complexly shaped mechanical parts has

also led to the possibility of using the ultra precision diamond machining in the

reformation of the optomechanics for small-sized, robust and high performance

optical systems. However, reflecting freeform surfaces machined directly into metal

parts are quite difficult to polish, and therefore their utilisation has so far been limited

to NIR and IR wavelengths. [1] Yet, ideally freeform mirror optics can provide

excellent performance especially in broad wavelength range applications, because

mirrors, unlike lenses, are free from chromatic aberrations. For this reason it is

important to be able to utilize freeform mirror optics also in UV-VIS wavelength

range. In addition, silicon dioxide (SiO2) or magnesium difluoride (MgF2) coatings

are used to protect aluminium mirrors. SiO2 coating offers protection to surface while

maintaining high reflectivity characteristics in UV/VIS region [2-4]. To test the level

of diamond machining and coating four freeform mirrors were manufactured in this

study. Micro milling of freeform mirrors opens up new possibilities for making UV-

VIS components.

2 Manufacturing and Coating

Two sets containing two identical freeform mirrors from MS358 naval brass and

Alumec 89 tooling aluminum were pre-machine from an Ironcad based STEP file by

multi-axis high speed machining. The STEP file was then transferred to

Pro/Engineering program for ultra precision machining programming. To evaluate the

quality of STEP based programming additional reference lens cavity of diameter

13 mm, 40 mm radius and 0.5 mm sag was created in Power Shape program. The lens

file was imported as an STEP element to the Pro/E system and merged with the

freeform mirror file. In Pro/ E a three axis raster micro milling tool path with 0.1 µm

tolerance was generated for cutting the freeform mirror and the reference lens. Since

freeform machined parts are fairly difficult to measure it was decided that the

reference lens would be used for assessing the quality of the diamond machining.

Milling of the parts was then done on a Moore 350 FG diamond machine tool with a

controlled waviness 2.5 mm radius diamond milling tool from A.L.M.T corporation.

After the machining of the aluminium and brass mirrors, a 100 nm SiO2 coating was

applied by using Beneq TFS 200 atomic layer deposition (ALD) device.

36


3 Characterisation and Spectral Testing

The freeform aluminium mirror with and without SiO2 coatings was measured with

WYKO NT9300 optical profiler. Surface roughness Ra on uncoated aluminium

mirrors reference lens was 6.8 nm and for uncoated brass mirrors reference lens the

Ra was 5.7 nm. On the SiO2 coated aluminium mirror surface roughness Ra was

5.2 nm and on coated brass mirror Ra was 5.5 nm.

The test mirrors were originally designed to be used as probes with a fiber optic

spectrometer. Therefore test measurements were carried out with the same setup

(shown in Fig. 1a)): the aluminium and brass mirrors were connected (one at a time)

to a spectrometer (MultiSpec, TEC5 AG, Oberursel, Germany) with UV-VIS optical

fibers for both illumination and collection.

a) b)

Figure 1: a) Measurement setup with the fiber optic mirror probe. b) Change in

reflectivity of the test mirrors after applying the ALD coating.

The overall reflectivity was significantly higher (up to about 3x) with the aluminium

mirrors at 250-500 nm wavelength range, but at longer wavelengths the reflectivities

were almost identical. The two brass mirrors showed a larger mutual difference in

reflectivity, whereas the two aluminium mirrors behaved identically. The SiO2 ALD

coating lowered the reflectivity of all mirrors throughout the measured UV-VIS

spectrum. The brass mirrors showed a larger spectral variance in the drop of

reflectivity (see Fig. 1b)). It must be kept in mind that the purpose of applying a

coating is not in enhancing reflectivity, but smoothing the surface roughness, as well

as protecting the surface from environmental degradation.

37


4 Conclusions

In this paper two aluminium and brass freeform mirrors were ultra precision micro-

milled and SiO2 coating was added to enhance the optical characteristics of the

mirrors in UV/VIS region. The after-milling optical characteristics of the uncoated

and SiO2 coated mirrors were analysed and the spectral functionality of the mirror

systems tested by reflectance standards. The results show that Ra values of micro-

milled reference lenses on uncoated aluminium and brass surfaces were in good

correlation to theoretical values and slight improvement in these values can be seen

with coated aluminium and brass surfaces. The spectral measurements show that SiO2

coating affects the reflectance characteristics of both mirror surfaces while

maintaining high reflectivity characteristics.

Acknowledgement

The work in this paper was supported by TEKES/European Union-European Re-

gional Development Fund.

References:

[1] K.Kataja, M. Aikio, K. Niemelä, and M. Aikio, "Optimization of Free-Form

Illumination Optics", Key Engineering Materials 364-366, 724-727 (2007).

[2] A. Gebhardt, S.Scheiding, "Manufacturing of Freeforms with well- defined

Reference Structures", in OptoNet Workshop – Ultra precision manufacturing of

Aspheres and freeforms Jena, Fraunhofer Institute for Applied Optics and Precision

Engineering IOF, Sept. 22-23, (2010).

[3] P. J. Smilie, B. S. Dutterer, J. L. Lineberger, M. A. Davies, and T. J. Suleski,

"Freeform Micromachining of an Infrared Alvarez Lens" in Advanced Fabrication

Technologies for Micro/Nano Optics and Photonics IV, W. V. Schoenfeld,

J. J. Wang, M. Loncar, T. J. Suleski, eds. Proc. of SPIE Vol. 7927, 79270K (2011).

[4] X. Jiang, P. Scott, and D. Whitehouse, "Freeform Surface Characterisation - A

Fresh Strategy", Annals of the CIRP 56, 553-556, (2011).

38


Photonic Flip-Flop based Solutions to overcome

Memory-Wall Challenges

Paul Tcheg1, Bei Wang1, Merih Palandöken1 and Tolga Tekin1,2 (1) Forschungsschwerpunkt Technologien der Mikroperipherik, TU-Berlin (2)Fraunhofer-Institut für Zuverlässigkeit und Mikrointegration (IZM)

[email protected]

Abstract

In this paper, a compact-designed hybrid-integrated all-optical flip-flop (AOFF), with

InP-based semiconductor optical amplifiers (SOA) integrated on planar silicon on

insulator (SOI) waveguide platform, is shown. For this purpose, all needed passive

components namely straight waveguide, s-bends und multimode interference (MMI)

couplers with s-bended waveguide, are investigated in the first part of report. The

Mach Zehnder interferometer (MZI) [1] is taken into account regarding the transfer

function for signal access for AOFF both with asymmetrical MMI and with

symmetrical MMI. A thermo-optical analysis is carried through with quasi-analytical

approach and simulations. In the second part, the compact-designed hybrid-integrated

AOFF is evaluated as system and its outputs are analyzed. It is shown that by

injecting 200 ps optical pulse train with 8.45 dBm power through its set and reset port

the AOFF changes its states. The proposed study is of interest in the design of

compact heterogeneous integrated AOFF.

1 Introduction

In order to satisfy the insatiable demand for HPCs used for server or for

entertainment like game computing, it is important to develop balanced computer.

This corresponds to a uniform growth in the performance both of CPUs and of

memory cells. From 1986 to 2000 the CPU-speed increased by 55% yearly while the

memory cell’s speed increased by 10% [2]. And the gap between the CPUs and

memory cell’s speed is called “memory wall” [3]. It is therefore to consider the

development of all-optical computer architecture; particularly AOFF. Considerable

investigations have been done concerning AOFF [4-7]. Due to the reliability of

CMOS-compability of highly photonic integrated circuits (PIC) the SOI-platform

seems to be up to now the most promising integration platform enabling the

realization of optical memory cells. Moreover SOI provides not only the optimal

39


mixture or adhesive strength of silicon with III-V materials, which dominate the area

of fast optical switches. The SOI-platform enables also the implementation of very

compact photonic circuits. This property was used to design the passive optical com-

ponents include in the device. The AOFF contains active, passive optical

components and also micro-heaters for adjusting the phase.

2 Building blocks of all-optical flip-flop

The AOFF device contains silicon nanophotonic elements as waveguide, s bend,

MMI couplers (72/28 and 50/50) and SOA. A device schematic depicting the

building blocks and their positioning on the SOI device is shown in fig. 1. The desc-

ription of the device’s functionality is given in [5,7]. Since TE polarization is

considered, all passive SOI components rely on 220nm height Si Strip waveguides.

reset set

Figure 1 : Schematical diagram of the designed all-optical flip-flop.

Strip waveguide: In order to avoid enhanced power losses due to power coupled to

more than one modes, a single-mode (SM) operation is required. To meet these con-

ditions, the strip waveguide has follow dimensions: h= 220nm height and W= 400nm

width as SM structure for TE polarization (fig. 2a). The Si strip waveguide is

surrounded by spin on glass (SoG) and silicon dioxide (SiO2) respectively as upper

and as lower cladding with refractive index nSoG=1.37, hSoG=580nm, nSiO2=1.46 and

hSio2=2μm. Fig.2b depicts the mode profile of the strip waveguide for TE polarization.

Figure 2: Strip waveguide cross section (a), mode profile (b) and Cross section of

thermo-optical phase section (c).

40


S-Bends: Two bends, both with a radius of 5.4μm, were designed for injecting light

into the MMI couplers. In order to reduce the propagation losses, a waveguide with a

length of 1μm was inserted between the two bends. Here the efficiency is 88%.

MMI coupler: From the proposed AOFF structures [4,7] a 72:28- and a 50:50-MMI

coupler should be taked into account. The access waveguides are ideally positioned to

allow the splitting ratio 72:28 [8] and 50:50 [9]. Both couplers are 3μm wide and

however consider the design conditions of MMI in [10]. The beat-length is 17.83μm

and length of the 72:28 MMI and 50:50 MMI are 10.69μm and 26.74μm respectively.

To avoid excess losses and ripples in the interfaces between the access waveguides

and the MMI-section due to the mode-mismatch and the discontinuities regarding the

dimensions, the optimal width of the access waveguide connected to the MMI-section

is 750 nm. The splitting ratios of 72:28- and 50:50 MMI are 70.2:27 and 48.3:48.3.

Thermo-optical phase section: By means of micro-heaters, the phase difference due

to the electrical supply of SOAs with differents peak values of current is compensated

[11]. Here Titanium was chosed as material for the heaters. Fig.2c represents the MZI

cross-section used to apply the thermo-optical study, where hTi= 100nm, Wheater=

500nm, hSoG= 580nm are the thickness and the width of titanium and the height of

SoG repectively. The parameters ΔTπ/2 = 3.54°K; Pπ/2 = 4.53mW; RTi = 4.8kΩ;

I=0.97mA used in the analysis. Required power and temperature difference for the

phase change of π/2 are also considered.

3 Simulation of the all-optical flip-flop as system

The AOFF with 6 IOs is simulated based on the results of Si PIC. Set and reset pulse

trains and CW bias signals wavelengths set as λ1=1556nm and λ2=1559.57nm with

input powers of 3.01dBm and 2.78dBm, respectively. The InP-SOAs were biased at

ISOA1=120mA, ISOA2=76mA. 200 ps wide optical pulses with wavelength of λpulse=

1562.5nm and power of 8.45 dBm were injected into the MZI that was controlling the

operation state. The period of pulse trains were 2ns. The output power by each MZI

will be filtered. Thereby exhibits the rise time 50 ps. It can be seen that the switching

between flip-flop states every 2 ns occurs and thereby the extinction ratio between the

41


states stays stable. Fig. 3 depicts the layout of the compact-designed AOFF with a

length of 3.34 mm and a wide of 1.29 mm. It contains s-bends and grating coupler for

further optical packaging [12].

4 Conclusion

In this article, we show a compact-designed all-optical flip-flop as photonic

integrated circuit on planar silicon on insulator (SOI) waveguide platform to address

the Memory Wall challenges of today’s HPC architecture. An overview on the design

steps of the building blocks rely on 220 nm height Si Strip waveguides was also

reported. Then it was shown that by injecting 200ps wide optical pulses with

8.45dBm power through its set and reset ports the AOFFs state changed dynamically.

Figure 3 : Layout of all-optical flip-flop with the coupling components.

References:

[1] T. Tekin et al., Proc. ECOC, pp. 123-124, 2000.

[2]S. McKee, Proceedings of the 1st conference on Computing, 2004.

[3] W.A. Wulf, et al. SIGARCH Comput. Archit. News, 23(1), 1995, pp. 20–24.

[4] Martin T. Hill, et al. Microw. Opt. Tecnol. Lett., 2001, 31, pp. 411-415.

[5] Martin T. Hill, et al. optics Letters / Vol. 30, N°13/ July 1, 2005.

[6] F. Ramos, et al. J. Lightwave Technol., 23, 2005, pp. 2993-3011.

[7] Y. Liu, et al. Electronics Letters vol. 42 N°. 24, 23rd Nov. 2006.

[8] T. Grzegorczyk, et al. 1997 Proceedings of ECIO 97, pp. 150-153. [9] Trung-Thanh Le, e-ISBN 978-3642-32183-2, IFMBE Proccedings.

[10] L.B. Soldano, et al. IEEE JLT, Vol.13, No.4. April 1995, pp. 615-627.

[11] Leuthold, et al. IEEE JQE 1998, Vol.34, Issue 4, pp.622-633.

[12] T. Tekin IEEE JSTQE vol. 17 (3),pp. 704-719, 2011.

42


High Precision Injection Moulding of Freeform Optics with

3D Error Compensation Strategy

L. Dick1,2, S. Risse3, A. Tünnermann2,3

1 JENOPTIK Polymer Systems GmbH, Germany 2 Friedrich Schiller University Jena, Abbe Center of Photonics, Institute of

Applied Physics, Germany 3 Fraunhofer Institute for Applied Optics and Precision Engineering IOF,

Germany

[email protected]

Abstract

Injection moulding offers a cost efficient method for manufacturing high precision

plastic optics in high volumes. In connection with the demand for freeform optics in

imaging optical systems like head mounted devices or head up displays [1],

unsymmetrical shrinkage compensation strategies have to be developed to realize

freeform optical surfaces with high precision for high volume applications.

This paper describes an efficient method for significantly increasing the form

accuracy of injection moulded freeform optics. In this regard, a typical plastic

freeform optics has been designed, moulded, and commonly optimized by main

moulding process parameters. The process-related shrinkage of the freeform optics

generated a non-rotationally symmetric surface error. To compensate for such kind of

non-uniform shrinkage, a freeform error surface had to be superimposed to the

freeform design surface on the mould. Regarding the measurement analysis, two

strategies are discussed. The first method is a best-fit procedure and in the second

case, well defined reference structures are used. In conclusion, the systematic form

deviation can successfully be pushed from the typical range of illumination optics

into the level of some imaging applications at moulded plastic optics.

1 Demonstrator design, mould- and process optimization

For analysing the process chain, a typical freeform optics was defined and

specifically modified for the processing with injection moulding. The demonstrator

design, the mould design as well as the cavity manufacturing process are shown in

figure 1. The freeform surface is described by a Zernike polynomial function and has

43



a clear aperture of 44 mm with a part diameter of 70 mm. The reference structures are

defined by three spheres (3 x 120°) with a convex radius of 10 mm on a reference

circle of 58 mm.

Figure 1: Lens design, mould design, and mould machining with Slow Tool Servo

In order to achieve sub-µm accuracies of smaller than 0.5 µm p-v at the mould, a

procedure already shown in [2] was used. Based on the high precision mould,

optimizations at the moulding process were done. So at least the main influence

parameters (melt- and mould temperatures, dwell pressure, dwell pressure- and

cooling times) [3] were systematically modified separate with respect to the form

deviation. Afterwards, a 33 experimental design was realised. Optimal process

parameters were defined and checked with FEM simulation methods. Main results of

a 33 experimental design are shown in figure 2. An optimal process was found at 1000

bar dwell pressure and 240°C melt temperature at a mould temperature of 90°C. The

process scatter seems to be a random, but for the defined process optimal as well.

Figure 2: Main influence moulding parameters on form deviation and process scatter

2 3D error compensation based on the best-fit strategy

Moulding the freeform optics with optimal process parameters, a median form

deviation of the moulded parts was calculated, taking into account measurements of

the 2½D profilometer of Panasonic UA3P. The measured points cloud was tilted and

shifted in all 6 degrees of freedom by using the least square method. In order to

44


compensate the 3D shrinkage of the moulded freeform surface, the form deviations

were averaged and superimposed to the mould design surface at right locations.

Further parts were moulded with the same process parameters. In conclusion, the

form deviation was reduced from 18.2 µm p-v / 4.29 µm rms to 1.57 µm p-v /0.25

µm rms.

Figure 3: Form deviation before (left) and after (middle / right) one iteration loop

3 3D error compensation based on reference structures

In order to determine the form- and position errors [4], the moulded freeform surface

and the three reference spheres were measured in one setup at the 2½D profilometer.

Subsequently, a new coordinate system based on the 3 reference points was build to

fix all 6 degrees of freedom. The measured points cloud was compared with the

mathematical description. The initial error based on this process is shown in figure

4, left. Results seem to be similar to the measured data of the best-fit deviation in

figure 3, left. The reason is that the references are realised high precisely in one

setup with the freeform surface itself. Furthermore, the shrinkage of the moulded

part is very symmetric, so that the reference points are able to shrink

homogeneously, directed to the centre. After the first iteration loop, a deviation map

with about 6.0 µm p-v / 1.1 µm rms was measured. Compared to other investigations

of this batch, the error map did not appear systematically and thus, further iterations

based on this method were not useful. The fact is caused by existing measurement

uncertainties and process scatters at more elements on the part.

By using an additional best-fit procedure on the error map, final tilt errors of 0,0056°

around the X and -0,0029° around the Y axis as existing main position error in this

case was detected. In consideration of this, a final form deviation of 1.65 µm p-v /

0.28 µm rms can be calculated.

45


Figure 4: Form- and position error before (left) and after (middle) the first iteration

loop, and resulting form deviation after an additional best fit procedure (right)

4 Summary

Due to several process parameters, moulding freeform optics lead to an influence of

the form deviation. Finding an optimal combination, usually high systematically

errors can be measured by using different analysis strategies. The asymmetric

systematic deviation map can be superimposed to the design surface at the mould.

Depending on the measurement analysis method, accuracies of less than 2 µm p-v

and of about 6 µm p-v were reached successfully on a demonstrator surface by using

the best fit strategy and by using the reference marks, respectively. For this, optimal

process parameters provided only about 20 µm.

The fundamental investigations were funded by the German Federal Ministry of

Education and Research (BMBF) within the project “FREE” – grant number

JENOPTIK Polymer Systems GmbH: 13N10826 and Fraunhofer IOF: 13N10827.

References:

[1] Eberhardt, R. in: „Freiformoptik – Die Herausforderung für zukünftige optische

Systeme“, LASER + PHOTONIK, 3 / 2010, 2010

[2] Dick, L. et al.: “Injection moulded high precision freeform optics for high

volume applications”, Advanced Optical Technologies Vol. 1, 2012

[3] Nievelstein W. in: „Die Verarbeitungsschwindung thermoplastischer

Formmassen.“, university Aachen, PhD thesis, 1984, Aachen, Germany

[4] Scheiding, S. et al.: “References – A Key Issue for Freeform Structuring”,

EUSPEN Special Interest Group Meeting. Structured and Freeform Surfaces Topical

Meeting, 10.-11.02.2010, Aachen, Germany

46


Integration Platform of Dual Wavelength Signal Source for

120GHz Wireless Communication Systems

Merih Palandöken1, Tolga Tekin1, 2 1Technische Universität Berlin, Germany

2 Fraunhofer-Institut für Zuverlässigkeit und Mikrointegration (IZM, , Germany)

[email protected]

Abstract

Monolithically integrated photonic signal sources at subterahertz frequencies are

becoming an attractive and compact solution for the future wireless communication

systems. An optical packaging of dual wavelength DFB laser is presented with the

assembly steps required for the optimum optical coupling, RF modulation and DC

biasing with optimum wiring circuitry in the housing, and better thermal

management while preserving the mechanical stability of housing. The laminate

based integration platform to be designed for the various modulating inputs in

addition to direct modulation input and active section of laser such as phase shifter

and SOAs are illustrated. The additional metallic parts required for better

mechanical stability and efficient heat removal during laser operation and high

temperature assembly steps are utilized in the packing process. The glass blocks for

the optimum fiber positioning in the optical coupling are also the important parts in

the assembly process to be highlighted. The whole customized package is illustrated

as an example of reliable laser packaging.

1 Introduction

Optical communication has been increasing its importance and presence from

backbone to access and premise applications in spite of the recent market slump.

DWDM is definitely an epoch-making breakthrough in the industry and promisingly

introduced to metropolitan area network. It's now very crucial to reduce the cost and

the size of light sources to penetrate deeper into the practical use [1]. Very compact

wavelength-tunable optical transmitter modules are therefore important system

components in metro WDM applications. This reality results the reliable source

packaging of compact transmitters to be an important task. A directly modulated

DFB chip can be the choice as an optical source for the cost reason.

47



In this paper, the customized packaging of currently developed module is explained

in combination with additional submounts to optimize DC and RF wiring circuitry

and respective assembly steps in the packaged module.

2 Laser Source Packaging Design Issues

The main design issue in monolithic signal source packaging is to have the required

DC and RF contacts with possibly small wire lengths and large separation distance

inbetween due to small inductance and capacitance for reduced RF coupling. The

optimum contact positioning and dimensioning are also important not to have

additional reflections at RF ports for high SNR and optical modulation efficiency.

Especially for the proper operation of active SOA/EAM and laser sections, the

resulting heat has to be removed effectively from chip during data modulation. In

addition to the operation point drift and linearity degradation of active components,

the dimensional change of optical waveguides in the passive components leads

mode profile to be degraded with possible multimode operation due to larger cross-

section. Therefore, not only RF and DC transmission lines have to be positioned

and geometrical parameters have to be determined in optimum manner, but also the

temperature gradient resulting from active components has to be minimized

throughout the submount for high data rate modulation. The geometrical parameters

and electrical properties of laser chip are shown in Figure 1 with the laser geometry.

Figure 1: DFB laser chip

48


3 Monolithically integrated Dual Wavelength Signal Source Package Design

In order to protect the laser chip from unwanted environmental effects to sustain

EMC between each components of the system, and especially for the

mechanical/thermal stability and to supply the required I/Os for the external

connections as test and control pins, the packaging design is quite important. To

ensure the optimal performance of the whole monolithic signal source system, the

following requirements must be fulfilled:

1-) Thermal and Thermo-Mechanical Requirements:

To ensure optimal thermal flow by avoiding thermo-mechanical mismatch due to

different thermal expansion coefficients of materials of assembled components and

smoothing occurring temperature gradients inside the laser chip are necessary for

optimal positioning and assembling of the different components and whole system

into the package. For the thermal analysis, the main heat sources are the active and

passive optical components with the respective loss powers as indicated in Figure 1.

Total thermal power to be removed from the chip is 2.6W. Therefore, a

thermoelectric module has to be used in combination with a high thermally

conductive material such as brass block as a support material underneath.

2-)Optical and Electrical Requirements :

There is one optical access in the target housing design to satisfy an optical access

on the left hand side of monolithic DFB laser source. Optimal fiber- coupling

through accurate and stable optical alignment is important along with DC and RF-

input contacts to feed the active optical components and modulating baseband data

in optimal manner. Therefore, two precisely micro machined glass blocks are used

to manage accurate optical alignment for the tilted optical input on the chip left-hand

side. Due to 23° tilt of optical input for low optical reflection at the chip edge, new

laminate based submounts have to be designed to accommodate this angle with

appropriate wiring distribution. These submounts are shown in Figure 2.a.

3-) I/O count (DC and RF):

There are totally one RF (SMAs) and 12 DC input accesses resulting from the

additional submounts designed in the final packaged monolithic laser module.

49


Furthermore 4 pins are required for the biasing and controlling TEC including NTC-

Thermistor.

The customized housing and respective package is shown in Figure 2.b.

(a) (b)

Figure 2: (a) iDWSS laser chip with respective DC and RF wiring submounts

(b) Customized housing and laser chip package

4 Conclusion

In this paper, the customized optical package of dual wavelength DFB laser

is explained with the important design parameters to be taken into account for

reliable laser packaging. The optimum optical coupling, laminate based RF

modulation and DC biasing submounts with optimum wiring circuitry, thermal

management with metallic support material and TEC in a customized housing

design are illustrated as an optical integration platform.

References:

[1] Harufi Yoneda, et al.,” A Compact 2.5Gbps Wavelength-Tunable DWDM

Transmitter with Direct-Modulated DFB”, Electronic Components and Technology

Conference, 2002

50

Session 1: Precision Engineering Advancements Enabling Progress in

Energy Technologies

51


Precision Engineering for Concentrating Solar Power (CSP)

Applications

C Sansom1, P Comley1, P King1, and N Macerol1

1Cranfield University, UK

[email protected]

Abstract

This paper describes the application of precision engineering principles and

techniques to CSP (Concentrating Solar Power) component manufacturing and

optical surface characterization. We explain the opportunities for precision engineers

to play an expanding role in the development of CSP technologies, by detailing the

current research within the Precision Engineering Institute at Cranfield University,

UK. This includes the characterization of large glass solar collectors using

photogrammetry and a large Coordinate Measuring Machine (CMM), the evaluation

of both glass and metallised polymer films for use as heliostats and parabolic

concentrators, and the simulated ageing of both glass and polymer film solar

collectors in hostile environments. We also discuss surface coatings to create self-

cleaning and anti-soiling surfaces.

1 Introduction

Concentrating Solar Power (CSP) uses large area solar collectors to concentrate direct

sunlight (DNI or Direct Normal Insolation) to a focal point or line, achieving

concentration ratios of up to 1:1500. The thermal energy is absorbed at the focus and

transported by a heat transfer fluid (HTF) to either a thermal energy storage tank or a

steam generator. Thereafter the destination of the energy depends on the application,

which includes electrical power generation, heating, cooling via steam enabled

chillers, water desalination, the provision of industrial process heat, water purification

and cooking. The technology differs fundamentally from solar PV (Photovoltaic) by

virtue of its integration of energy storage and the dispatchability of the energy

produced. In addition to many small and community scale installations there are

nearly 200 CSP power plants worldwide with an average output of 72 MW, an

example of which is the Andasol 50MW plant in southern Spain shown in Figure 1.

53



Figure 1a: Andasol CSP plant, Spain Figure 1b: Parabolic trough collector

2 Measurement of surface form

Cranfield University has developed a photogrammetry technique [1-2] for the on-site

measurement of CSP collectors, using a Canon EOS 18MPix DSLR camera with

PhotoModeler software plus in-house written visual display code and maps. Precision

is equivalent to less than 1/5 pixel RMS ~ 1:20000 with 18MPix, which translates to

~50 µm over a 1m object. The technique is validated by cross-reference to the

Cranfield University CMM which can accommodate optical reflectors of up to 3m x

2m x 1m. For a dimension of 1m, a CMM maximum permissible error of length

measurement of less than 5μm is achievable. Form measurements of a solar parabolic

collector section (see Figure 2a) and a 4m length absorber receiver tube

Figure 2a: 1.6m collector on CMM Figure 2b: Robot abrasive ageing

have been performed. Figure 3 shows a representative error map to illustrate the

deviation in form from the parabolic shape, extracted using the CMM. There is an

obvious twist in the panel with the rear-right corner being around 2mm higher than

the rear-left corner. The RMS error for this fit is 0.388mm

54


Error in x (mm)

Err

or

in z

(m

m)

Figure 3: Solar collector section showing deviation from parabolic shape

3 Optical surface characterization

Experiments to investigate the effect of cleaning processes on the optical surfaces of

collectors have been performed. The schedule for the contact cleaning experiments is

shown in Figure 4 below.

Figure 4: Contact cleaning experiments with sample designations

55


This work involved the use of a FANUC Robot M-710i (see Figure 2b) to simulate

abrasion by silica particles of varying geometries and hardness upon samples of both

1mm thick collector glass and silvered polymer film. An example of the optical

results obtained are shown in the summary graph of Total Reflectance in Figure 5.

Figure 5: Total Reflectance measurements of solar collector surfaces

The results indicate that the polymer film collectors compare favourably with the

glass collector samples when undergoing the contact cleaning processes that are

typically employed in solar power plants.

4 Conclusions

Measurements of the surface form of glass solar thermal collectors have been

performed by the use of an in-house photogrammetry technique and a large CMM.

The optical reflectance of both glass and metallised polymer film collector pieces has

been evaluated under the simulated conditions that represent contact and non-contact

cleaning processes. Surface form measurements illustrate the deviation of collector

form from the parabolic shape. Polymer film based collectors are shown to be robust

when subjected to the same cleaning processes as that experienced by glass collectors

in currently operational solar power plants.

5 References

[1] K. Pottler, E. Lüpfert, G. H. Johnston, M. R. Shortis: Photogrammetry: A

powerful tool for geometric analysis of solar concentrators and their components,

Journal of Solar Energy Engineering, Vol 127, pp. 94-101, February 2005.

[2 . lme . ei . le . e l e eas eme s a ab lic

Troughs Using the Reflected Image of the Absorber Tube, Journal of Solar Energy

Engineering, Vol 131, pp. 011014, February 2009.

56


Photocatalytic Activity Influenced By Thickness of TiO2

Measured in Nano and Macro Scale

S. Daviðsdóttir1 K. Dirscherl2, S. Canulescu1, R. Shabadi, R. Ambata1

1DTU and Denmark 2DFM and Denmark

[email protected]

Abstract

Titanium dioxide (TiO2) in the anatase crystalline structure corresponds to one of the

most powerful photocatalytic materials available today. Photons with the energy

equal (UV region) to or higher than its band gap (~3.2 e.V) are able to initiate a photo

activation process in TiO2, which creates hole/electrons pairs in the material. The

hole/electron pair consists of high oxidizing and reduction power respectively which

can be used for spliting water into hydroxyl radicals and converting oxygen into

superoxide.

The main focus of this paper is to map the photocatalytic activity at the nanometre

scale on the TiO2 coating with different thickness on aluminium substrate

synthesized by magnetron sputtering. Scanning Kelvin Probe Force Microscopy

(SKPFM) was used for nano-scale mapping of the photocatalytic activity under UV

light, while macroscopic electrochemical measurements were used as

complimentary method to compare the properties. Further, diffuse reflectance

measurements were used to determine the band gap as a function of thickness of the

coating. The three different techniques are correlating, Mapping the surface

potential in Nano scale revealed that the surface potential for thin films was less

homogenous, indicating an influence from a substrate oxide at the junction of the

coating and the substrate. The existence of the substrate oxide was detected by the

use of Transmission Electron Microscopy.

1 Introduction

Photocatalytic behaviour of TiO2 coatings have been investigated widely due to their

various potential applications for purification of water and air, as self-cleaning and

hydrophilic surface, and as antimicrobial surfaces for health care. The ability of

TiO2 coatings in the anatase phase to utilize the energy of light to decompose

57


compounds is well-known and has been reported in various investigations [1].

Electromagnetic waves with energies equal or higher to the band gap of TiO2, 3.2

eV, for anatase phase are able to generate electron-hole pair in the valence and

conduction band respectively. These electron–hole pairs can either migrate to the

surface and participate in redox/oxidation reactions of the dirt or recombine within

the coating. The oxidation power of the holes is sufficient to form hydroxyl radicals

that are the key to the photocatalytic degradation ability of TiO2. Moreover, the

reducing ability of the electrons can form superoxide.

The effect of the coating thickness of TiO2 on glass substrate on photocatalytic

activity has been reported in several publications [2]. However, the study of TiO2 of

various films thickness on metallic substrate is rare.

2 Materials and method

2.2. Sample preparation

The substrate materials used for the present investigation were standard AA1050

aluminium alloy. Prior to the coating, aluminium specimens were polished to 1

micron surface finish using a buffing machine (Polette 6NE from KE MOTOR A/S).

The coating synthesis was carried out by pulsed DC magnetron sputtering using an

industrial CemeCon CC800/9 SinOx coating unit. The nominal film thicknesses

ranging from 100 nm to 2 µm.

2.3. Characterisation technique

2.2.1. Scanning Kelvin Probe Force Microscopy

The SKPFM instrument used for the investigation was “Multimode V” (Bruker).

Scanning of the surface was carried out in interleave mode in which the tip scans the

topography first followed by surface potential scanning by lifting the tip by 100 nm,

which is kept constant.

2.2.2. Electrochemical measurements

For the electrochemical measurements, a standard three electrode flat

electrochemical cell set-up was used. In the flat cell, the specimen was held against

an O-ring exposing a surface area of 9.6 cm2 to the solution. The reference electrode

used was Hg/Hg2SO4/saturated K2SO4 in order to avoid any chloride contamination

58


in the test solution. The electrolyte used for all the experiments was 0.1M NaNO3

(AR grade) dissolved in de-ionized water as a supporting electrolyte with pH 5.5.

The volume of the electrolyte filled in the cell was 550 ml. The UV lamp used was

Philips original home solarium.

2.2.3. Optical measurements

Optical diffuse reflectance measurements were carried out to investigate the UV

absorption characteristics of the TiO2 coating, which in turn was used to determine

the band gap of the semiconductor. The diffuse reflectance of UV-visible spectra are

converted to equivalent absorption using the Kubelka–Munk model [3]. The

diffused reflectance measurements were carried out using an optical setup consists

of an integrating sphere system (reflectance geometry 8°/d).

3 Results and conclusion

When the hole-electron pairs are formed by photo excitation, the surface potential

changes. The hole will be consumed by the electrolyte and the electron will pass to

the substrate. The absolute change of the surface potential is highly dependent on the

energy of the photon generated electrons. That is how much photon energy the

electrons require in order to enter to the conduction band.

Figure 1. surface potential with and without UV light on TiO2 films with thicknesses

of A) 100 nm B) 500 nm and C) 2 µm

The change of local surface potential using SKPFM is presented in Figure 1 by

overlaying the potential values on the topography image. The colour bars indicate

the potential values in Voltage. The left side of each picture show SKPFM image of

the films without UV illumination, while the right side shows corresponding images

of films under UV illumination. The figures show clearly that upon illumination,

surface potential changes. However there is no change in topography when the

No UV With UV No UV With UV No UV With UV

A C B

1µm 1µm

0.2µm 1µm

1µm

0.2µm 1µm

1µm

0.2µm

59


sample is illuminated. The sample with a thickness of 100 nm shows an

inhomogeneous surface potential distribution, while with increasing thickness the

surface potential distribution becomes uniform. The uniform distribution indicates

that the coating is activated uniformly with no non-active sites and that the synthesis

method by sputtering process produces an anatase structure at Nano-scale.

In order to compare the nano-scale activation of the coating measured by SKPFM

with overall surface activation, conventional open circuit potential measurement was

used with and without UV light exposure. Moreover, the poetical shift was

compared to the band-gap of the sample obtained by optical measurements. The

comparison of the methods can be seen in figure 2

-0.2

-0.15

33.13.23.33.43.5

150

200

250

300

350

400

Band gap [eV]Excitation potential, OCP [V]

Pote

ntial shift,

SK

PF

M [

V]

Figure 2: Comparison of three different techniques and their correlation. The surface

potential shift when the sample is excited by UV-illumination, measured by SKPFM

technique. The OCP measurement occurred when the sample is under illumination.

The band gap is obtained from reflectance measurements.

References:

[1] a Fujishima, T. Rao, and D. Tryk, “Titanium dioxide photocatalysis,”

Journal of Photochemistry and Photobiology C: Photochemistry Reviews,

vol. 1, no. 1, pp. 1–21, Jun. 2000.

[2] P. Activity and I. Introduction, “Dependence of TiO 2 Photocatalytic

Activity upon Its Film Thickness,” no. 21, pp. 360–364, 1997.

[3] I. Tunc, M. Bruns, H. Gliemann, M. Grunze, and P. Koelsch, “Bandgap

determination and charge separation in Ag@TiO2 core shell nanoparticle

films,” Surface and Interface Analysis, vol. 42, no. 6–7, pp. 835–841, May

2010.

100 nm

500 nm

2 µm

60


An Unconventional Experimental Setup for Testing Cutting

Performance/ Wear Resistance of Diamond Cutting Wires

V. Herold1, S. König1, M. Berg2 1University Jena, Institute for Materials Science and Technology, Germany 2j-fiber GmbH, Germany

[email protected]

Abstract

Diamond wire cutting (fixed abrasive wire cutting) is an economically promising

alternative technology for slurry-based loose abrasive wire cutting in the field of

slicing of hard crystalline materials, especially for photovoltaic Si-wafers. Main

function-related characteristics for a practical application in wafer manufacturing are

both cutting performance and wear resistance. The achieved procedure and the

experimental setup permit the testing of the diamond cutting wires under conditions

comparable to the real slicing process, but with drastically reduced consumption of

work material and diamond cutting wire.

1 Introduction

Manufacturing quality and economic efficiency in the production process of

crystalline photovoltaic Si-wafers are dominated by a reliable control of the slicing

process. For fixed abrasive wire cutting the tool characteristics, determined by the

basic wire (material, diameter, cross sectional shape), the abrasive grains and their

distribution (diamond type, grain size, grain shape, grain distances, grain bonding

overstanding, depth of the grain embedding), as well as their bonding characteristics

(bonding material, thickness of the bonding) are significant for the cutting

performance to be obtained. [1] The main criteria used for the evaluation of diamond

cutting wires are the cutting performance or the abrasive properties (material removal

rate), the surface quality of the wafers (geometrical parameters, surface roughness,

subsurface damage) and the wear resistance (tool’s lifetime). The cutting parameters

(cutting speed, feed speed and reverse factor) can be optimised, adjusted to a present

type of cutting wire and depending on the machinability of the material to cut.

61



2 Measuring methods and experimental setup

2.1 Characterization of the abrasive layer of the diamond cutting wires

Diamond cutting wires consist of a basic wire with diamond grains, which are in most

cases fixed by an electroplated bond. In fact the diamond cutting wire is a three

dimensional entity, the positions of the cutting edges of the abrasive grains can be

described in cylindrical coordinates (x (axial), r (radial), (circumferential)). For

reasons of simplification the measurement and description of the grinding layer were

made two-dimensional, i. e. the positions of the abrasive grains cutting edges are

given in Cartesian coordinates in a plane which includes the wire axis. Parameters of

the abrasive layer of grinding tools (cutting edge positions, geometrical parameters of

the grains) are random variables, which can be described by appropriate distribution

functions. For investigation of the geometrical properties of abrasive tools both

qualitative methods (light microscope, SEM) and quantitative methods (optical

scanning or tactile scanning with a profilometer) are established. [2] In the latter case

function-related characteristics (distribution of static cutting edges, distribution of

distances between cutting edges) can be calculated.In the framework of these

investigations the profiles of the diamond cutting wires were measured by optical

scanning (CNC coordinate measuring machine VideoCheck / Werth Messtechnik

GmbH) and by tactile profilometry with a knife edge probe (FormTalysurf /

AMETEK Taylor Hobson Ltd.)

2.2 Experimental setup for testing of the cutting performance with

measurement of cutting forces and wire bow

The experimental setup is shown in Fig. 1. The translational cutting motion in the real

wafer slicing process is replaced by a rotational cutting motion, which is performed

by a disc-shaped Si-test specimen. In addition there is a component of relative motion

along a programmed path and with intermittent feed after every loop. In maximum 8

single equally-tensioned wires are arranged in a wire frame. The measuring systems

for cutting force/ feed force and for wire bow are integrated in the base plate of the

wire frame. The experimental setup was realized on the base of a CNC surface

grinding machine Planomat 408 (Blohm).

62


Figure 1: Experimental setup for the simulation of the cutting process

3 Experimental results

The SEM-photographs of different diamond cutting wires (Figure 2) illustrate varying

abrasive layers with different grain distributions. It is obvious, that these cutting wires

should present divergent cutting performance and wear resistance.

Figure 2: SEM-photographs of two different types of diamond cutting wire

Figure 3: Profile of the same section of the diamond cutting wire in different wear

states (measurement by tactile profilometry) ; wire type JC, vc = 10 m/s,

vf = 0,5 mm/min, monocrystalline silicon

The profiles in Figure 3 document the suitability of the experimental setup for

detailed wear investigations. The special construction of the wire clamping avoids

torsion of the diamond cutting wires so that dedicated sections can be refound in the

63


profile after use in the cutting process. The wire bow during the cutting process is

related to the feed force and to changes in the cutting performance of the diamond

cutting wires. The end of the tool’s life is announced by deterioration of the cutting

performance and an increasing wire bow. The wire bow in cutting of monocrystalline

silicon is lower as compared to the cutting of multicrystalline silicon according to the

expectations (Figure 4).

Figure 4: Wire bow for different removal rates and Si-materials types to cut

4 Conclusion

An unconventional testing system for simulation of the cutting process with special

kinematics with integrated measuring systems for the cutting force / feed force as

well as wire bow was developed for comparative investigations of different types of

cutting wires. Thus it is possible to analyse interrelations between the topography of

the abrasive layer of the diamond cutting wires, their cutting performance and wear

resistance with minimized consumption of work material and diamond cutting wire.

Because of the special construction of the experimental setup the same section of the

diamond cutting wire can be measured before starting and after several grinding

cycles. Therefore the wear can be observed at dedicated diamond grains.

References:

[1] Development of precision fixed diamond wire PWS; Nakaruma, N.;

Kazahaya, K. et al: DIAMOND TOOLING JOURNAL 03(2011) p. 28–31

[2] Final report concerning CIRP cooperative work on characterization of

grinding wheel topography; Verkerk, J.; Delft 1977

64


Fabrication of Freeformed Blazed Gratings by

Ultraprescision Machining

K. Haskic1*, S. Kühne1*, S. Lemke2, M. Schmidt1 1Technische Universität Berlin, Institut Für Werkzeugmaschinen und Fabrikbetrieb

(IWF), Fachgebiet Mikro- und Feingeräte (MFG), Germany 2Helmholtz-Zentrum Berlin für Materialien und Energie (HZB), Institut

Nanometeroptik und Technologie (G-INT), Germany *Equally contributing

[email protected]

Abstract

There are a few methods to produce high quality plane symmetric gratings. But until

now it was only possible to produce blazed gratings with so called grating machines.

This is a mechanical process and the grooves are divided with a special formed

diamond tools in thin metal layers. Because the material is reallocated during this

process it is difficult to produce coarse gratings without deformation of the structure.

Hereinafter, a method is presented to produce blazed gratings with a planning

process. The structures are not deformed and there are no limitations to substrate

forms.

1 Produced structures

With this process it is possible to produce differed types of gratings. The blaze angle

is not restricted and even echelle gratings with angles up to 80° and densities down to

15 grooves per millimetre were produced. On the other hand also typical echelette

gratings with line densities up to 300 grooves per millimetre and blaze angles down

to 3° were produced. The substrate form is not limited and gratings were planned in

spherical (concave and convex) substrates. Different types of substrate forms are also

possible. The radius of curvature and the maximum differences in height are not

restricted. Gratings with a radius of curvature between 8 and 300 mm were already

produced. The blaze angle can be changed for every groove. So it is possible that the

grating normal is always perpendicular to the substrate surface. It has been also

shown that it is possible to fabricate two or more different areas of blaze angles in

one grating without a border between the areas (fig. 1).

65



Figure 1:AFM measurement, grating with two different blaze angles (3° and 15°)

2 Used machine

A modified and improved ultra precision milling machine LT-Ultra MMC-1100

(fig. 3) [1] was used to fabricate all of the gratings. The tilt and rotation module (fig.

2) [1] adds two additional rotation axes which allow changing the angle of the

diamond planning tool directly during fabrication.

Figure 2: Tilt and rotation Module Figure 3: LT-Ultra MMC-1100

The base structure of granite reduces vibration and temperature effects. Additionally

an active air damping system isolates the machine from the ground to achieve very

low roughness values on the gratings. A high stability of the groove density is

achieved with the linear motors controlled by corrected glass scales. The used

correction algorithm will be described in further publications. In combination with

the hydrostatic bearings there is no stick slip and a start-stop-motion of the slow axis

is used for fabrication.

2.1 Tool alignment

Synthetic monocrystalline planning diamonds were used for fabrication. With an

adjustment device it is possible to bring one of the tool corners direct in the rotation

axis of the tilt and rotation module. The adjustment is done with very fine adjustment

screws on solid state joints and monitored with a high resolution video capturing. An

66


inbuilt vacuum chuck is used to fixate the tool after adjustment. This procedure is

essential because if the corner is not exactly in the rotation centre the lower blaze

edge will shift over the grating length.

3 Fabrication process

The material is removed via planning/cutting from the substrate. The position and

angle of the tool is NC-controlled. To achieve high quality blaze surfaces a multi-cut-

process was used. Dependent on structure geometries several pre-cut cycles are

needed to remove material with no measurable wear to the diamond (fig. 5). The last

finishing step (fig. 6) and the correct choice of the offsets select the blaze facet

(echelette or echelle). If this step is missing it is impossible to produce echelle

gratings because the chip is ripping on the steep blaze and cut on the antiblaze. This

step is not required for echelette gratings but it improves the roughness. A schematic

visualisation of the cutting steps is shown in fig. 4 for an echelle grating with a total

of 4 cuts. The gratings were machined into electroplated gold with the use of

lubricant. The surface structure of the gold is of no importance because it will be

removed and only the homogeneity of the gold must meet higher claims. The use of

lubricant reduced the measured planning forces down to 60% and influences the chip

behaviour positive.

Figire 4: Multi-cut process Figure 5: Pre-cut Figure 6: Finishing

(1-4: sequence of cuts)

4 Parameters, “ghosts” and quality

The fabrication paramters directly influence the rougness and thus the quality of the

blaze. The chip surface and the feed influence the force on the diamond and the wear.

The smallest burst on the cutting edge will increase the stray light. The cutting forces

were measured to identify parameter sets with low wear and good surface quality but

will be presented in further publications in more detail. A small infeed between 2 µm

1

2

3

4

67


for echelette and 4 µm for echelle gratings with a feed greater 1000 mm/min gives the

best results and an area roughness below 2 nm was realized. The stray light was

measured in the littrow-blaze condition with a HeNe-Laser. The signal period of the

glass scale is too small to produce rowland ghosts [2] but lyman ghosts [3] still occur

if no correction is used. This is an alternative to the typical used laser interferometers

first presented by Harrision [4] to control the slow axis. In figure 7 are examples of a

stray light measurement with a satellite caused by non periodical errors due to

material inhomogeneities and a stray light measurement of a high quality grating with

an almost ideal order. The stray light of the high quality grating in this example is

below 0.0001 at the used wavelength.

Figure 7: Stray light measurements, left) with satellite, right) ideal order

5 Summary

It has been shown that it is possible to produce high quality plane and spherical

gratings. The multi-cut-process reduces the roughness of the blaze down too 2 nm

and even ghost and other errors don’t appear. Additionally the blaze angel can be

changed for every groove and the structures are not deformed. Thus planning is an

adequate process for grating fabrication.

References:

[1] LT-Ultra Precision Technology GmbH, Aftholderberg, Germany

[2] R.W. Wood (1924). Phil.Mag. Ser. 6, Vol. 48, Issue 285, pp. 497-508

[3] T. Lyman (1901). Proc.Am.Acad.Arts Sci., Vol. 36 No. 14, pp 241-252

[4] G.R. Harrison, J.E. Archer (1951). JOSA, Vol. 41, Issue 8, pp. 495-502

68

Session 2: Nano & Micro Metrology

69


In-line metrology of functional surfaces with a focus on

defect assessment on large area Roll to Roll substrates

L. Blunt1, L Fleming1, M. Elrawemi1, D. Robbins2, H. Muhamedsalih1 1 University of Huddersfield, UK, 2 Centre for Process Innovation, Sedgefield, UK

[email protected]

Abstract

This paper reports on the recent work carried out as part of the initial stages of the

EU funded NanoMend project. The project seeks to develop integrated process

inspection, cleaning, repair for nano-scale thin films on large area substrates.

Flexible photovoltaic (PV) films based on CIGS (Copper Indium Gallium Selenide

CuInxGa(1-x)Se2) have been reported to have light energy conversion efficiencies as

high as 19%. CIGS based multi-layer flexible devices are fabricated on polymer film

by the repeated deposition, and patterning, of thin layer materials using roll-to-roll

processes (R2R), where the whole film is approximately 3µm thick prior to final

encapsulation. The resultant films are lightweight and easily adaptable to building

integration. Current wide scale implementation however is hampered by long term

degradation of efficiency due to water ingress to the CIGS modules causing

electrical shorts and efficiency drops. The present work reports on the use of areal

surface metrology to correlate defect morphology with water vapour transmission

rate (WVTR) through the protective barrier coatings.

1 Introduction

To address the PV degradation problem a thin (~40nm) coating of Al2O3 has been

implemented to provide the environmental protection (barrier) for the PV cells. The

highly conformal aluminium oxide barrier layer is produced by atomic layer

deposition (ALD) onto a Polyethylene naphthalate (PEN) substrate. The surface of

the starting polymer is further planarised to give a high quality smooth surface prior

to ALD. The presence of surface defects, pin holes and debris particles on the

uncoated film can create significant defects within the, aluminium oxide, barrier

layer. This paper reports the results of measurements conducted to characterise the

uncoated and barrier coated polymer film surface topography using segmentation

71


feature parameter analysis. The presence of defects is then correlated with the water

vapour transmission rate as measured on representative sets of films using a standard

MOCON test. The paper also shows initial measurement taken on a prototype in

process, high speed, environmentally robust optical interferometer instrument

developed to detect defects on the polymer film during manufacture. These results

provide the basis for the development of R2R in process metrology devices for

defect detection

2 Barrier Substrate

A series of 4 coated substrates were produced having a 40nm ALD Al2O3 barrier

coating. An area of 80mm2

was used for testing of the

barrier properties using a

standard MOCON test, fig 1.

This test measures the steady

state WVTR for a barrier

coating under defined

conditions. The system

places the substrate in a sealed unit where one side of the substrate is subect to high

humidity and the other side is defined as the dry side. The dry side is purged with a

carrier gas which carries away any transmitted water vapour to a infrared sensor

which records the transmission rate. The steady state rate was recorded along with the

time to stable transmission.

3 Results

The WVTR results show that

sample 2705 had a significantly

higher WVTR than the other

specimens. Following WVTR

testing the surface topography of

all samples (including uncoated)

was measured using laboratory based coherance correlation interferometery.

Sample No WVTR (g/m²/24 hrs.) Time

2701 1.1x 10-3 11 days

2702 1.3 x 10-3 11 days

2705 4.1x 10-3 5 days

2706 2.0x 10-3 5 days

Figure 1: MOCON test set up [1]

Table: 1 Water vapor transmission rate at

specified conditions 38o

C and 90% RH

72


700 measurements, equating to 14% of the total

surface area of all the specimens was

measured. The results showed the presence of

defects both particulate and pin hole type in all

specimens. Typical examples are shown in fig

2. The surface roughness of defect free

samples was ~0.6nm. Areal topography

characterisation was caried out using the

feature parameter set ISO 25178-pt2. In

particular the parameter Sfd was used (where

Sfd = the number of significant hills +

significant dales); significance was defined as

any peak/pit greater than 20% of the total peak

to valley roughness (Sz). Using this default

significance value for the defects the results

showed no clear correlation with the WVTR

results. However when the significance critera

was increased for hills (Peaks) and dales(pits)

(over +/- 3xRMS roughness and >15µm lateral

dimension) the Sfd paramter could be used to

count only the most severe defects over the total

measured area. In this case the correlation was

clear (figure 3b). The results indicate the

presence of small numbers of large defects

dominate the WVTR of the barrier layer. ALD

coating is highly conformal and is likely to coat

particulate debris and down deep pits. The

mechanism for increased WVTR would be that

debris on the surface or within pits become

detached exposing uncoated PEN to water

ingress.

Figure 2 Typical substrate defects a) large

hole 60µm wide, 385x383µm b)

particulate debris 30nm high, 113 x

113um c) White light scanning

interferometry meaured defect, 153

x121µm.

Scale 2.65 µm

Scale 5.5 µm

Scale 40nm

73


Figure 3a defect density (all defects) b) Significant defect count

The aim of these results is to implement on line metrology for the roll to roll ALD

process and using the knowledge gained from the present work implement areal

feature analysis to carry in-line metrology and process control. The above analysis is

now combined with a robust wavelength scanning intferometery instrument having

internal environmental compensation to carry out the measurement work using

parrallel sensors to cover large areas of the substrate surface.

Conclusions

The present work has shown the potential of areal feature segmentation to detect

functionally significant defects present in roll to roll produced ALD barrier coatings

of Al2O3. The results show a good correlation between the presence of small

numbers of large defects and WVTR. The analysis provides the basis for in process

metrology for roll 2 roll production of barrier coatings for flexible PV arrays and is a

first demonstration of in process use of feature parameters. Work is continuing to

check repeatabiliy of these tests and produce “cleaner” substrates.

Acknowledgement

The authors would like to thank the EC for providing funds to carry out this work

via the NanoMend project NMP4 LA-2011-280581.

References:

[1] Duncan, B. Urquhart, J and Roberts, S. (2005). Review of Measurement and

Modelling of Permeation and Diffusion in Polymers. NPL Report DEPCR 012

[2] X. Jiang, K Wang, F. Gao, and H. Muhamedsalih “Fast surface measurement

using wavelength scanning interferometry with compensation of environmental

noise” Applied Optics May 2010, Vol. 49, No. 15.pp/ 2903

[3] ISO 25178-pt2 ( 2012) Surface texture: Areal -- Part 2: Terms, definitions and

surface texture parameters

74


High-resolution Investigation and Application of

Diamond-coated Probing Spheres for CMM- and Form

Metrology

M. Neugebauer, S. Bütefisch, T. Dziomba, S. Koslowski, H. Reimann

Physikalisch-Technische Bundesanstalt (PTB), Bundesallee 100, 38116

Braunschweig, Germany

[email protected]

Abstract

Diamond-coated probing spheres with diameters from 0.3 mm to 8 mm to be used in

form and coordinate metrology were investigated and first results are presented. The

diamond-coated probing spheres have a very smooth surface, low form deviations

and are resistant against wear.

1 Introduction

In coordinate and form metrology, workpieces of different materials are mostly

measured by tactile probing. Contacting elements are usually spheres made of ruby,

sapphire or ceramics like alumina, zirkonia or silicon nitride. Depending on the mate-

rial probed and on the contact pressure, mechanical wear or adherence may occur on

the surface of the probing sphere. To avoid these effects, diamond probing spheres

can be used. However, diamond spheres are expensive and not easy to obtain in all

diameters needed. Moreover, diamond spheres cannot be manufactured perfectly

spherically because of their crystal structure. Probing spheres coated with poly-crystal

synthetic nano-crystalline diamond [1, 2] were developed to avoid the above

mentioned problems. We investigated the geometrical properties, the surface

characteristics and the probing behaviour of diamond-coated spheres ( 8 mm down

to 0.3 mm) in comparison to standard ruby spheres ( 8 mm down to 1 mm).

2 Characteristics of the diamond-coated probing spheres

Figure 1 shows a number of styli with diamond-coated spheres [3]. On the left-hand

side, different CMM styli are shown and on the right-hand side, a special stylus can

be seen which is to be mounted onto a micro probe to be used with a micro CMM.

75


Figure 1: CCM styli made of tungsten carbide with diamond-coated probing spheres,

8 mm to 0.3 mm.

2.1 Surface texture

The surface texture as well as wear effects and adherence effects were investigated

using a white light interferometer with a magnification of 50x1 (WLI). Figure 2, left,

shows the surface texture after a spherical fit of a 5 mm diamond-coated sphere

(Sz = 8 nm, Sq = 1.1 nm) and, right, of a 5 mm ruby sphere (Sz = 37 nm,

Sq = 2.3 nm), FOV 80 µm x 80 µm, respectively. The diamond-coated surface is

mostly homogeneous and relatively smooth, compared with the ruby surface.

Figure 2: Surface of a diamond-coated sphere (left) and a ruby sphere (right).

2.2 Geometry

The form deviations of the probing spheres were measured both at the equatorial

zone and over the zenith, with the aid of a micro CMM F25, in comparison to

reference spheres. With the exception of the 8 mm probing sphere, the form

76


deviations of the diamond-coated probing spheres are about 0.1 µm, or even smaller.

Therefore, these probing spheres are well suited for geometrical measurements.

2.3 Mechanical wear

The wear was determined at a steel ring gauge using 1 mm probing spheres. The

ground surface of this unique ring coincidentally embeds a number of grinding

particles. The measurements were carried out with a form tester MFU8 (F 20 mN).

Three roundness profiles and four straightness profiles were measured five times

each as is usual in calibration work. The overall distance measured was about 11 m.

The wear at the ruby sphere amounts to about 0.9 µm and is shown in figure 3. No

wear was detected at the diamond-coated sphere as well as at the steel ring gauge.

Figure 3: Mechanical wear at the 1 mm ruby sphere; measurements with WLI,

FOV 150 µm (left) and detailed view with FOV 50 µm (right).

2.4 Adherence of the material probed

The affinity to adhere to material at the sphere’s surface during probing was tested at

a cylinder made of duralumin which has a relatively rough surface. The measure-

ments were carried out with a 1 mm diamond-coated sphere and with a 1 mm

ruby sphere, using a form tester MFU110 (F 20 mN). Five roundness profiles and

five straightness profiles were measured over a distance of about 0.5 m. At both

probes, duralumin adhered to the surface as shown in figure 4. The amount of

adherence to the diamond-coated sphere is less than that to the ruby sphere but it is,

nevertheless, not negligible for high-precision geometrical measurements. The

adherence could not be eliminated by mechanical cleaning but could completely be

eliminated at both spheres using Aluminium Etchant (ANPE 80/5/5/10).

77


Figure 4: Adherence of duralumin to the 1 mm diamond-coated sphere (left) and

to the 1 mm ruby sphere (right); measurements with WLI, FOV 150 µm.

3 Probing behaviour

With the form tester, a number of objects are measured: a test cylinder with a high-

quality surface, a 30 mm ceramic sphere and two 10 mm test cylinders which

have relatively rough surfaces and are made of bronze and duralumin, respectively.

The standard deviations obtained with the diamond-coated spheres were up to 50 %

less than those obtained with the ruby spheres. However, due to the adherence

effect, the duralumin cylinder could not be measured reasonably with both probes.

4 Conclusion and outlook

In a first step, diamond-coated probing spheres were investigated with respect to their

geometry and surface texture. Their probing behaviour was tested with the aid of a

form tester. In a second step, these probing spheres will be tested with a CMM

Prismo and, mounted onto micro probes, with a micro CMM F25.

5 Acknowledgement

The work was partially funded by Carl Zeiss Industrielle Messtechnik GmbH. The

authors gratefully acknowledge the outstanding cooperation of Mr. Wim Nelissen [3].

References:

[1] Balmer R. S. et al: 2009 J. Phys.: Condens. Matter 21 364221

[2] Nelissen W.: Mikroniek 2012 (Vol. 52) 3, 22-25

[3] http://www.diamondproductsolutions.nl

78


Validation of On-machine Microfeatures Volume

Measurement Using Micro EDM Milling Tool Electrode as

Touch Probe

G. Tristo1, M. Balcon1, S. Carmignato2, G. Bissacco3 1 Department of Industrial Engineering, University of Padua, Italy 2 Department of Management and Engineering, University of Padua, Italy 3 Department of Mechanical Engineering, Technical University of Denmark,

Denmark

[email protected]

Abstract

In micro electrical discharge milling, process parameters have to be empirically

calibrated in order to achieve high precision machining; to this end, on-machine

measurement of the material removed is of paramount importance. The capability of

electrical discharge machines in detecting electrical contacts between the electrodes

can be exploited to perform dimensional measurements, using the tool electrode

similarly to the touch probe in a coordinate measuring machine. In this work an

investigation of the accuracy of the on-the-machine volume measurements in a micro

electrical discharge milling setup is carried out and an evaluation of the error

affecting on-machine measurements is provided.

1 Introduction

Micro electrical discharge milling (µEDM milling) is a particular configuration of

µEDM where material removal is achieved exploiting electrical discharges occurring

between two electrodes and microfeatures are fabricated driving a cylindrical tool

electrode along tool paths as in conventional milling operations [1].

Since material removal and tool wear rates are strongly dependent on specific

working conditions, it is necessary to calibrate process parameters before proper

machining in order to produce high precision micro features. Accurate determination

of the amount of material removed from both tool and workpiece is thus of

paramount importance. To this end, on-machine volume measurements are needed,

especially to implement self-learning procedures for process parameters optimization.

79



In commercially available EDM machines it is possible to exploit the short-circuits

detection system that is used for the electrical discharge machining process also to

obtain dimensional measurements. In previous works, the short-circuit detection

system has been adopted for roundness deviation evaluation [2] and the repeatability

and reliability of tool length measurements performed with this method has been

assessed [3]. However the use of the short-circuit detection system and the tool

electrode to perform coordinate measurements similarly to a touch probe in a CMM

have not been reported yet and a metrological validation of this micro EDM

measuring method is missing [4].

2 Errors induced by imperfections in the machined geometry

On-machine volume measurements based on coordinate measurements by the tool

electrode are mainly influenced by the dimensional measurement capability of the

µEDM system and by the most relevant imperfections present in machined features,

such as surface roughness, walls draft angle and corners rounding. Experiments were

performed on a Sarix SX-200 µEDM machine. The probe used for on-machine

measurements was fabricated with SX-200 wire dress unit and characterized with the

optical sensor of a Werth Video-Check-IP 400 multisensor CMM.

A circular pocket with a diameter of about 515 µm and a depth of about 440 µm was

machined on a block of mould steel by µEDM milling using a 300 µm tungsten

carbide tool electrode and a finishing set of process parameters. Then the workpiece

was cross sectioned in correspondence to the centre axis of the hole. SEM images

(figure 1-A) and confocal measurements show that floor and wall surfaces have

comparable but non negligible surface roughness. As a consequence, when the

cylindrical tool-probe is used to measure the diameter and depth of the cavity it

touches the burrs around the craters instead of the average profile of surfaces,

producing a systematic under-estimation of the quantity of material removed during

the erosion of the pocket. The volume per unit of surface was estimated and it was

evaluated that up to 0.7% less volume was measured because of surface roughness.

The corners rounding radius on the floor of the pocket (figure 1-B/C) was measured

with the optical CMM (measured radius: 45 µm); then the related over-estimation of

the pocket volume was evaluated to be within 1%. The angle of inclination of the

walls of the pocket (figure 1-C) was measured with the optical CMM to be about 0.7

80


degrees; as a result, in the worst case the associated volume measurement error for

these specific geometries is 3%.

Figure 1: SEM image of the tool-probe used for on-machine measurements (A). SEM

image (B) and micrograph (C) of a cross sectioned blind hole machined by µEDM.

Distribution of measured points on: (E) a step specimen, obtained assembling a

calibrated gauge block having a nominal height of 100 µm on a larger flat surface,

and (D) a through hole with a diameter of about 500 µm machined by µEDM.

3 Determination of dimensional measurements uncertainty

The uncertainty of dimensional measurements performed by the µEDM milling

machine was determined following the experimental method standardized in ISO

15530-3 [5], through 20 repeated measurements of calibrated workpieces. To this

end, the diameter of a calibrated through-hole was measured on the µEDM machine,

acquiring 30 points equally spaced along the circumference as in figure 1-D, while

the height of a calibrated step specimen was measured as in figure 2-E. Reference

calibrations of the diameter and height were performed using the Werth multisensor

CMM mentioned before.

The experiments showed that the expanded uncertainty (coverage probability of 95%)

for depth measurements is equal to 1.3 µm and the standard deviation of the 20

repetitions is 0.26 µm, while for diameter measurements the expanded uncertainty is

81


equal to 1.9 µm and the standard deviation is 0.5 µm, after the correction of a

systematic error, quantified in 3.1 µm.

4 Discussion and conclusions

Analysing the results, the most relevant quantity influencing the accuracy of on-

machine volume measurements is the slope of the walls. This measurement error

contribution depends not only on the extent of the walls slope, which is proportional

to the aspect ratio, but also on the depth at which the diameter of the pocket is

measured. Theoretically it is possible to calculate the exact depth where to measure

the diameter to nullify the measurement error associated to wall slope: given the

small value of the draft angle, this level can be approximated to half of the hole

depth. Corners rounding radius and surface roughness, instead, are constant for a

given set of process parameters; hence the associated errors are mainly dependent on

the volume and surface-to-volume ratio of the measured cavity. The measurement

method showed a good repeatability, but a significant systematic error was found in

the diameter measurement. Correcting this systematic error allows measurement

uncertainties below 2 µm (coverage factor k=2).

In conclusion, this work showed that it is possible to perform on-machine volume

measurements with relative errors below 3%, which is acceptable for calibration of

process parameters.

References:

[1] K. Ho, et al., “State of the art electrical discharge machining (EDM)”,

International Journal of Machine Tools and Manufacture, 43, 1287–1300, 2003.

[2] D.-Y. Sheu, “Study on an evaluation method of micro CMM spherical

stylus tips by µ-EDM on-machine measurement,” Journal of Micromechanics and

Microengineering, 20, 075003, 2010.

[3] G. Bissacco, G. Tristo, and J. Valentincic, “Assessment of Electrode Wear

Measurement in Micro EDM Milling”, in Proceedings of the 7th International

Conference on Multi-Material Micro Manufacture, 155–158, 2010.

[4] S. Carmignato, et al., “Traceable volume measurements using coordinate

measuring systems,” CIRP Annals - Manufacturing Technology, 60, 519–522, 2011.

[5] ISO 15530-3: 2011, “Geometrical Product Specifications (GPS) –

Coordinate Measuring Machines (CMM): Technique for Determining the Uncertainty

of Measurement – Part 3: Use of Calibrated Workpieces or Measurement Standards”.

International Organization for Standardization. Genève, 2011.

82


Virtual CMM method applied to aspherical lens parameters

calibration

A. Küng, A. Nicolet and F. MeliFederal Institute of Metrology [email protected]

Abstract

Micro coordinate measuring machines (µ-CMMs) are attractive to accurately measure

optical components like aspheres. Nevertheless, providing the measurement

uncertainty for each parameter of the asphere is not trivial. Therefore, a parametric

fitting algorithm coupled with a virtual µ-CMM based on a realistic model of the

machine was developed to perform Monte Carlo simulations and provide the asphere

parameters uncertainties.

1 Introduction

Tactile ultra-precise coordinate measuring machines such as the METAS µ-CMM

(fig. 1) are commonly used for measuring

optical components. This instrument

exhibits a single point measurement

uncertainty in the range of a few

nanometres, even in scanning mode [1],

which renders it very attractive for

measuring optical components having high

slopes like aspheres. Nevertheless, the

analytic estimation of the measurement

uncertainty for each asphere parameter is

almost impossible because of the many

combined influences from the measurement

strategy, the applied fitting procedures and

the different sensitivities of each one of the asphere parameters. The application of a

Monte Carlo method offers a simple solution to this complex problem.

The example described in this paper can be easily used for the calibration of any

complex parametric surface other than an asphere.

Figure 1: The METAS µ-CMM.

83


2 Parametric form fitting

Parametric form fitting in 3D is not that simple since the fitting algorithm must

minimise the sum of least square distances between the measured points and the

parametric surface by iterating on the surface parameters. Hence, these distances must

be computed orthogonally to the parametric surface, which also requires an iterative

process in each point. For the parametric form fit we used the Levenberg-Marquardt

iterative algorithm. Since this algorithm relies on the local derivative of the

parameters to converge to a local minimum, one has to pay attention to the initial

guess parameters, the stopping criteria, the relative sensitivity between the parameters

and the symmetries of the parametric form.

2.1 Asphere fitting

In our specific case of fitting an asphere, the fitting parameters are Xc, Yc and Zc the

coordinates of the asphere centre, X and Y the rotation angles around the X, and Y

axis (Z can be eliminated as an asphere has a rotation symmetry along the Z axis),

R the asphere radius, K the asphere conicity and C2, C4, C6 ... the asphere parameters

as given by the generic asphere equation:

=/ଶݎ

1 + ඥ1 − (1 + ଶ/ଶݎ(ܭ+ ݎଶܥ

ଶ + ݎସܥସ + ݎܥ

+ ⋯ where: =ݎ √ଶ + ଶ

Since the C2, C4, C6... coefficients are multiplying r2, r4, r6..., they do not have the

same relative weight. In order to re-equilibrate their weight, parameters C2, C4, C6...

were replaced by C'2 10-2, C'4 10-4, C'6 10-6,.... so the incremental step in the fitting

algorithm is thus more or less equally weighted for each parameter, and guaranties

the stability of the fit convergence to a local minimum solution.

For a robust fitting, the initial guess parameters where computed by fitting a sphere to

the measured points, whereas all other remaining initial guess parameters can be

chosen to be zero. The stopping condition was set to the least significant digit of the

computer in order to insure a good fitting even for the highest order Cx parameter.

2.2 Implementation

The fitting algorithm was implemented in Labview, as an iterative algorithm. The

Quindos software forwards the coordinates of the measured points and the initial

guess parameters for the asphere to the Labview executable by means of a file. The

84


Labview executable replies back to Quindos by sending the coefficients of the fitted

asphere (fig. 2).

3 Monte Carlo simulation to provide measurement uncertainty

The fitting algorithm of the 3D parametric surface provides information about the

quality of the fitting through the covariance matrix. Nevertheless, in order to provide

a measurement uncertainty for each parameter of the asphere, one has to include the

uncertainty of the measuring machine, the measurement strategy and measurement

conditions. Therefore a numerical model of the µCMM measurement process was

developed in which the error contributions were previously determined by

measurements. The model includes six basic types of contributions:

- Single point repeatability - Linear length variation

- Residual axis orthogonalities - Machine axis straightness

- Probing sphere residual shape error - Thermal drifts

This model can then be used to perform Monte Carlo simulations of a specific

measurement task [2].

3.1 Determining the asphere measurement uncertainty

First, a real measurement is performed on the aspheric artefact. Then, to each data

point from the asphere measurement, a simulated measurement variation is added

using the realistic numerical model of the µ-CMM. An asphere is then again fitted to

all these newly simulated points using the algorithm described in paragraph 2, and

Figure 2: Asphere artefact under measurement and residual form deviation afterfitting of less than ±50 nm.

85


the asphere parameters as well as the

covariance matrix are stored. This virtual

measurement is then repeated many

hundred times to finally deliver statistical

data for each asphere parameter (fig. 3).

3.2 Results

The many results of the virtual

measurements are processed with a

separate software. The statistical

evaluation in the table here below reveals an uncertainty which includes the

measurement strategy, the machine geometry errors, temperature drifts, etc.

ParameterX

(°)

Y

(°)

Xc

(mm)

Yc

(mm)

Zc

(mm)

R

(mm)

K C2

(mm-1)

C4

(mm-3)

C6

(mm-5)

C8

(mm-7)

C10

(mm-9)

value 0.0187 0.012 0.00272 0.00486 0.00038 9.849 -0.459 0.158 0.213 0.353 -0.3839 0.320

uncertainty 0.0006 0.007 0.00013 0.00012 0.00004 0.014 0.008 0.007 0.008 0.008 0.0049 0.007

In addition, the cross-correlations between the asphere parameters can be analysed,

for instance the strong correlation between the asphere radius R and the quadratic

parameter C2 in the asphere equation. This explains the large uncertainty of 14 µm on

R, even though the variations induced by the virtual CMM are smaller than 25 nm!

Fixing one of these two parameters ( C2 = 0 ) is usually more meaningful.

4 Conclusion

Any parametric surface such as an asphere can be calibrated. Thanks to the realistic

model of our µCMM, measurement uncertainties for each parameter can be delivered.

Additionally, the eventual cross-correlation between parameters can be analysed.

Acknowledgement:

This work was part of the EMRP IND10 Project. EMRP is jointly funded by the

EMRP participating countries within EURAMET and the European Union.

References:

[1] Proc. of the 7th euspen Int. Conf. – Bremen - May 2007, Vol. 1, 230-233

[2] Proc. of the 10th euspen Int. Conf. – Delft - June 2010, Vol. 1, 91-94

Figure 3: The virtual µ-CMM process

86


3D shape measurement under multiple refraction condition

using optical projection method

Yoshihisa Uchida, Ryota Kamei, Yuki Higashio

Department of Mechanical Engineering, Aichi Institute of Technology

1247 Yachigusa, Yakusa-cho, Toyota 470-0392, Japan

[email protected]

Abstract

We propose the improved 3D shape measurement system under multiple refraction

condition using optical projection method. This system consists of a laser, a camera

and a computer. This system using the optical projection method is a system that

projects a line pattern from the laser to the object surface, captures the object surface

image by the camera, processes the acquired image information with the computer by

the principle of the triangulation, and records and shows the shape of the object

surface on the display. New image analysis method is performed using Matlab for

calibration, calculation and display. We evaluated this system in air and water

conditions, experimentally. Results show that the 3D shape can be reconstructed

correctly. Experimental results also indicated that the average errors for X, Y and Z

are 0.05, 0.02 and 0.04 mm in both conditions, respectively.

1 Introduction

Three-dimensional (3D) shape measurement system has many applications such as

anthropometry in medical field and product inspection in industry. Therefore, many

measurement methods have been proposed by several researchers [1, 2]. However,

most of them measure only in air. In recent years, the 3D shape measurement system

has a wide industrial application for engineering from the micro-nano precision

measurement to the wide area measurement. Thus, the 3D shape measurement system

is needed in various conditions such as in gas, liquid and vacuum. Examples of

application of 3D shape measurement are work piece measurement in reactive gas

and liquid for micro-nano process, wear measurement of object in oil and precise

automatic control of robot arm in space. Therefore, it is important to develop the 3D

shape measurement system which can measure the object in medium of various

87



refractive indexes. Until now, 3D shape measurement systems in water using

ultrasonic have been proposed by researchers. These measurement systems can detect

only for wide measurement range.

We had developed the 3D shape measurement system and experimental results

showed that the 3D shape can be reconstructed correctly [3]. However, the system

accuracies are not enough for the precise shape measurement. In this paper, we

propose the improved 3D shape measurement system under multiple refraction

condition using an optical line projection method and investigate this system in detail.

New calibration and image analysis method is also proposed.

2 Measurement Principle

The schematic diagram of the 3D shape measurement system is shown in Fig.1. The

system consists of a semiconductor laser, a rotating mirror, a camera, a computer and

a display. A line pattern from the laser is projected on a surface of an object. The

incidence angle(α) is defined by the rotating mirror angle. The deformed pattern is

captured by the camera, which is set perpendicular to the line pattern direction. The

object, the laser with the mirror, and the camera form an optical triangulation system.

Therefore, the 3D physical spatial coordinates can converted from the 2D camera

image coordinates. The object is located in the various refractive indexes such as air

and water. In this case, the incident and reflect light beams travel through two or

more mediums. Thus, transformation equation is applied using Snell's law for various

refractive indexes. And we selected the line pattern projection method to obtain a

high optical intensity.

Figure 1: Schematic of measurement system

88


3 Measurement Procedure

Flowchart of the measurement system is shown in Fig.2. The measurement procedure

starts from the setting of the system including camera calibration. Parameters which

are the laser position, initial incidence angle, camera position, camera angle and

boundary are determined. However, there are various error causes such as the laser

position error, the incidence angle error, the camera position error, the camera angle

error, the distortion of lens aberration, etc. The known reference 50 points are pre-

measured to calibrate the system. 2nd step is projection processing of the pattern and

capture processing of the pattern on the

surface of the object. We used difference

image to improve an accuracy of the image

processing using background image. 3rd step

is image data processing. Erosion and dilation

are used to image denoising. To detect

maximum brightness for direction of

perpendicular to the line pattern, Gaussian

distribution function approximation is used.

Therefore, sub-pixel accuracy for Y direction

can be expected. 4th step is coordinate

transformation from 2D to 3D using

transformation equation. To measure the

whole object 3D data, we repeat the step 2-4.

And final step is storage and display of 3D

measurement data of the object surface. The

measurement process is performed using

Matlab.

4 Results and Discussion

In present system, basic experimental conditions are 1280 x 960pixel camera,

C(0,175,200), L(0,0,260), B(0,0,80), P(0,25,0)(object center) and n1=1.000 (in air).

Measurement area is X=225mm and Y=180mm at Z=0mm. Calculated resolution of

the camera for X, Y and Z are 0.18, 0.18 and 0.31mm, respectively. A sample object

is 50.5x50.5x20mm with 1-5mm trench, 1-5mm depth and white. To evaluate the

1st step

System setup and calibration

2nd step

Line pattern projection

and image capture

3rd step

Image data processing

4th step

Coordinate transformation

5th step

Storage and display

Figure 2: Flowchart of

measurement system

89


effect of the refractive index, n2=1.00 and 1.33 (in water, uniform condition) at 20oC

are selected. Number of measurements is 5. Figure 3 shows the measurement error as

a function of the true value for X, Y

and Z axes. Experimental results

indicated that the average errors for

X, Y and Z are 0.05, 0.02 and

0.04mm in both conditions,

respectively. Results also indicated

that the standard deviations for X, Y

and Z are 0.10, 0.07 and 0.09mm in

air and 0.14, 0.07 and 0.09mm in

water, respectively. These values are

under resolution of camera in present

system. However, some results have

errors larger than camera resolution

due to the edge effect of the captured

image.

5 Concluding Remarks

In this paper, we propose the

improved 3D shape measurement

system under multiple refraction

condition using optical line

projection method. Image analysis is

performed using Matlab for

calibration, image analysis,

calculation and display in short time.

References:

[1] K.Tsujioka, et al., Proc. SPIE, 7156, CD-ROM, 1-6 (2008).

[2] R. Menon and H.Smith, J. Opt. Soc. Am. A 23(9) 2290-2294 (2006).

[3] R.Kamei, et al., Proc. 12th euspen Int. Conf, Vol.1, 218-222 (2012)

-0.50

-0.25

0.00

0.25

0.50

0 2 4 6

Me

asu

rem

en

t er

ror[

mm

]True value [mm]

X-axisAir-Air

Air-Water

-0.50

-0.25

0.00

0.25

0.50

-20 -10 0 10 20

Me

asu

rem

en

t er

ror[

mm

]

True value [mm]

Y-axis

Air-Air

Air-Water

-0.50

-0.25

0.00

0.25

0.50

0 2 4 6

Me

asu

rem

ent

erro

r[m

m]

True value [mm]

Z-axis

Air-Air

Air-Water

Figure 3: Measurement error for X, Y, Z

90


Elastic behaviour of millimetre-scale polymeric triskelion-

like flexures

D.G. Chetywnd, Z. Davletzhanova, Y. Kogoshi, H. ur Rashid

School of Engineering, University of Warwick, UK

[email protected]

Abstract

A small study of the non-linear translational spring behaviour in triskelion planar

suspensions uses low-cost mm-scale polymeric devices to explore the effect of

several design parameters. The paper summarizes the approach and presents

illustrative results. The stiffening characteristic is often quite modest and the angles

of suspension beams may be useful for fine-tuning to different applications.

1 Introduction

The three-beam planar flexures called triskelions are attracting considerable interest.

E.g., some commercial CMM microprobes now exploit micro-fabricated triskelions

[1], to provide modest control of stiffness values over m-scale motions in three

translational freedoms. Other applications will benefit from different compromises

between potential operational parameters. Nano-force transfer standards [2] are an

excellent example, ideally needing very predictable z-stiffness in an approximation to

single-freedom translation. Taking a triskelion to be a symmetrical thin structure in

the xy plane, its central hub relatively easily undergoes small z-translations and x- and

y-axis rotations by means of bending and torsion of the suspension beams; the other

three freedoms are effectively constrained by the much higher stiffnesses on their

axes. Overall, it behaves similarly to a reduced-stiffness diaphragm. Intuitively, the

three suspension beams relate to the three freedoms, but even an elementary pseudo-

kinematic view of published designs indicates a significant over-constraint (mobility

well below three); so, e.g., non-linear (stiffening) spring behaviour is expected. The

present study therefore asks whether somewhat larger triskelions might be exploitable

as guides or reference springs in low-cost instrument systems, seeking practical data

on how design parameters might be selected to suit differing applications.

91



2 mm-scale polymeric triskelions

All published parametric models

seem to derive closely from [3], a

linear elastic model of symmetrical

forms having rigid platforms and

rigid inner arms with 60° ‘elbow’

connections (as in microprobes).

More complete models clearly

need experimental data for how

strongly spring non-linearity and relative axis stiffnesses vary with sizes and angles

within the triskelion flexure. Testing micro-devices is challenging and costly, so an

easier regime is desired. Making low-cost triskelions suggests injection moulding.

For both reasons, this study focusses on acrylic polymer devices at millimetre scales.

As the pilot study needed a few each of several design variants, specimens were hand

fabricated using simple open moulds machined into aluminium substrates. The

negatives of a relatively deep central hub and outer ring, connected by shallower leg

structures, were filled with a commercial acrylic surface replication resin and

smoothed off with a microscope slide. Despite the relatively poor control of this

manual process, beam thickness (the most critical dimension) repeated to within a

spread of 20 µm. Fig. 1 shows a device with 120° elbow angle. For all results given

here, the central hub radius was 1.5 mm and all beam sections were 1 mm wide. The

‘rigid’ arm and hub were both 1 mm deep. In alternate angled beam designs both leg

and arm had the same thickness. Having more modes of deflection (freedom), they

would be expected, if stable, to be less stiff and more linear than rigid-arm designs.

3 Measuring force-displacement characteristics

Figure 2 schematizes the force-displacement test-rig used. A side-acting inductive

gauge (T) carries a 50 mm probe arm (A) to which is clamped a small, hard

spherical probe tip and a saturated magnet that forms a force actuator with a

solenoid coil (FT). The tip contacts the hub of the specimen (S) held on a fine

motion xyz-stage. The gauge (Taylor Hobson Talymin) offers a range of 0.2 mm,

resolving practically to ~50 nm. The actuator provides a force closely proportional

to coil current up to about 1 N that is essentially independent of small variations in

Figure 1: 120° acrylic angle-beam triskelion

Arm

Hub

Leg

5 mm

92


magnet position, although there is a

small effect from the internal springs

of the sensor. A second displacement

gauge (H, an optical grating, 1 µm

resolution) monitors changes in the z-

height of the stage and sample and

allows a manual nulling technique.

The sample was raised by a set

amount and then the force increased to deflect the hub downwards to its previous

position (as indicated by T). This provides both larger range and reduced

uncertainty from internal springs.

First trials located the tip on the outer ring of specimens, acting rather like a micro-

hardness tester. Indentations were just discernible above the noise floor for loads up

to 500 mN, providing a low-quality estimate for Young’s modulus in the region of

1-5 GPa, reasonable for the polymer. This confirms that sample indentation is

negligible for stiffness measurements up to at least several kN m-1.

Summarizing (through limited space) results from rigid arm triskelion designs,

centrally loaded devices with suspension beams nominally 0.1-0.2 mm thick and

4 mm long could deflect by over 1 mm without failure. They had slowly stiffening

force-displacement curves that could consistently be fitted by a 3rd order polynomial

with R2 > 0.999; this is to be further investigated. Typically, stiffness was constant

within 1% up to over 200 µm deflection with a 60° elbow, but only to ~100 µm for

90° ones. Dimensions and materials properties were not closely controlled for valid

numerical comparisons, but short-range stiffness values were broadly consistent

with basic linear models and the patterns observed qualitatively as might be

expected from over-constraint on the internal bending and torsion modes.

Figure 3 shows results from a 60° angle beam device of 0.1 mm thickness. The

inner arm was 2 mm long, the outer leg 4 mm. The curve is noticeably straighter

than from a similar rigid arm design up to 1 mm. Initial stiffness is ~530 N m-1,

lower than rigid arm designs by less than might be expected intuitively. The longer

leg section dominates the arm, which it still allows some relaxation. The stiffnesses

for 90° and 120° elbows were around 1 kN m-1 and 700 N m-1. The 120° design was

the most non-linear, the 90° one the least. The reduced internal constraint might

T

XYZ

H

S

F

T A

Figure 2: Test-rig schematic

93


also reduce torsional stiffness at the hub to make the devices less useful as linear

springs. Applying the load at points close to the hub edge (the worst case for a real

application) showed that additional twisting reduced the stiffness under the load by a

consistent 7% at the reasonable limit for ‘linear’ behaviour. This suggests that the

same internal deflection modes dominate all the out-of-plane motions.

Figure 3: Typical stiffeneing behaviour for a 60 angle-beam triskelion

4 Conclusions

This pilot study encourages further investigations. While numerical comparisons

are unwise with current results, consistent patterns show these low-cost polymeric

triskelions have useful near-linear ranges within slowly stiffening characteristics.

There is clear scope for tuning performance by deviating from the ‘classic’ design.

References:

[1] e.g. IBS Precision Engineering ibspe.com/category/triskelion-touch-probes

accessed February 2013

[2] Pril W O 2002 PhD Thesis, University of Eindhoven

[3] Jones C W, Chetwynd D G, Singh J and Leach R K 2011 Proc. 11th euspen Int.

Conf., V1 191-194, Como

Acknowledgment

The authors are grateful to Prof. Richard Leach and Dr Chris Jones at the UK

National Physical Laboratory, whose large contributions and support of other work

in this field led directly to the present study.

94


Scanning results and repeatability testing of the TriNano

ultra precision CMM

A.J.M. Moers1, M.C.J.M. van Riel1,2, E.J.C. Bos1 1Xpress Precision Engineering, The Netherlands 2Eindhoven University of Technology, The Netherlands

[email protected]

Abstract

A novel coordinate measuring machine has been developed to provide a cost effective

solution for measuring micro components with a 3D uncertainty of 100 nm. This

paper summarizes the design aspects and part of the verification experiments

concerning repeatability and surface scanning using a Gannen XP 3D probing system.

Figure 1. Left side: TriNano CMM (artist impression), Right side: measurement using

the Gannen XM probe.

1 Operating principle

In the TriNano, the workpiece moves in three directions with respect to the stationary

probe by means of three identical linear translation stages. The stages are positioned

orthogonally and in parallel and support the workpiece table via vacuum preloaded

(VPL) porous air bearings as shown schematically in two dimensions in figure 2.

From this figure the operating principle of the TriNano becomes clear. A linear

translation of a stage is transferred via a VPL air bearing to the workpiece table.

Translations of the workpiece table with respect to the linear stage in other directions

than the translation of the stage are decoupled by the VPL air bearing. In this manner,

the three stages independently determine the position of the workpiece table in three

dimensions. On each linear stage, the scale of an optical linear encoder is mounted.

At the point of intersection of the measurement axes of these encoders, the probe tip

95


is located. As the orientation of the encoder scale does not vary with respect to the

probe, as can be seen in figure 2, the TriNano complies with the Abbe principle over

its entire measurement range. As a result, rotations of the workpiece table will have

little effect on the measured dimension.

Figure 2. Schematic representation of the operating principle.

Instead of a conventional orientation of the machine axes, i.e. two orthogonal axes in

the horizontal plane and a third vertically oriented axis, the three axes in the TriNano

are oriented such that each stage experiences an equal gravitational load. This

orientation of the axes combined with the operating principle results in identical

translation stages which can be produced at a lower cost.

This parallel configuration allows a low and equal actuated moving mass of each

stage with short and stiff structural loops. On machine measurements show that the

lowest natural frequency in the positioning loop is 75 Hz. This allows a high control

bandwidth, required for scanning measurements of micro parts with a velocity of 1-2

mm/s.

2 Thermal stability and compensation

Thermally induced errors are often the largest contribution to the total error budget in

precision measurement equipment [2,3,4]. However, certain straightforward measures

can be taken to reduce these thermally induced errors, such as minimizing and

controlling the heat flow and decreasing the thermal sensitivity of the machine. In the

Trinano, a pneumatic weight compensation system is applied to minimize the heat

production in the actuators. Furthermore, the relatively large granite frame results in a

long thermal time constant of the frame parts in the metrology loop. The other key

components in the metrology loop are the probe holder and the workpiece table.

Instead of applying a low expansion material like invar, these components are made

96


of aluminium and compensation for thermal expansion is implemented. The

temperature variations for the linear compensation model are measured by NTC’s

which are distributed in the metrology loop. The main advantage of NTC’s compared

to other sensors like PT100’s is their resolution [5] which is better than 0.1 mK.

-15

-10

-5

0

5

10

15

11 11.5 12 12.5 13 13.5 14 14.5

time in hours, start at 11:12

repeatability with temperature compensation

-150

-100

-50

0

50

11 11.5 12 12.5 13 13.5 14 14.5

rep

ea

tab

ility in

nm

time in hours, start at 11:12

repeatability without compensation

Figure 3. Left: Uncompensated single point repeatability values. Right: Single point

repeatability with compensation for thermal expansion.

Single point repeatability of the TriNano CMM using a Gannen XP 3D probing

system is verified on a steel gauge block. The results include the disturbances of all

parts of the metrology loop, e.g. thermally induced errors and the stability of the

vacuum preloaded air bearings. More information about the stability of the air gap of

the vacuum preloaded air bearings in the metrology loop has been published

previously [6].

Measurements show that, after compensation, a peak-to-valley deviation of 28 nm

over a 3 hour measurement can be obtained (without covers). The uncompensated and

compensated measurement results for single point repeatability are shown in figure 3.

3 Scanning measurements

To verify the dynamic behaviour of all components in the metrology loop, including a

Gannen XP probe, scanning tests are performed at a scanning velocity of 1 mm/s. The

measurement object is an optical flat which is scanned using a Gannen XP probe with

a ruby tip of 0.3 mm in diameter. The top surface of the optical flat is measured using

97


two square patterns, as shown in the left-hand graph of figure 4. From this graph it

can be seen that the optical flat was slightly tilted during the measurement.

contact points in mm

-15-10

-50

510

15

X in mm

-15-10

-50

510

15

Y in mm

4.97

4.98

4.99

5

5.01

5.02

5.03

scanning repeatability in nm

-15

-10

-5

0

5

10

15

X in mm

-15

-10

-5

0

5

10

15

Y in mm

-30-20-10

0

102030

Figure 4. left: Contact points scanning cycle. right: Repeatability of two subsequent

scanning cycles.

The same pattern on the optical flat was measured twice. The difference between all

corresponding measurement values without averaging of both scanning cycles is

within a band of ± 20 nm, as shown in the right-hand graph of figure 4.

4 Conclusions

Two important aspects which determine the performance of this CMM are stability

and the dynamic behaviour during scanning. After compensation for thermal

expansion, single point measurements show that the top-top deviation is within 28 nm

during a 3 hours period. The difference between repeated scanning cycles at 1 mm/s

is within a band of ± 20 nm.

References

[1] Rosielle, Constructieprincipes 1, Lecture notes 4007, Eindhoven University of

Technology, 2003

[2] Bryan, International Status of Thermal Error Research, Annals of the CIRP, 39/2,

1990

[3] Ramesh et al., Error compensation in machine tools - a review, Part II: thermal

errors, International Journal of Machine Tools & Manufacturing, 40, pp. 1257-1284,

2000

[4] Van den Bergh, Reducing Thermal Errors of CMM Located on the Shop-Floor,

PhD Thesis, Katholieke Universiteit Leuven, 2001

[5] Ruijl, Ultra Precision CMM, PhD thesis, Delft University of Technology, 2001

[6] Moers et al., Design and verification of the TriNano ultra precision CMM, IWK

Ilmenau, 2011

98


Distance ranging using original fiber-optic interferometer

K. Thurner, P.-F. Braun, K. Karrai

attocube systems AG, Königinstrasse 11a RGB, D-80538 München, Germany

[email protected]

Abstract

The requirements for accurate metrology in technologies such as extreme ultraviolet

lithography or compact machine tooling are continuously growing with the progress

in science and engineering. Such

applications often require extreme

environments like cryogenic

temperatures, ultra-high vacuum or

high magnetic fields and are often

constrained in space and volume. For

all these purposes we developed an

ultra-compact interferometer (figure 1)

based on the technique previously

reported in [1]. It is capable of

measuring displacements with nanometer repeatability and sub-nanometer resolution

for three axes at the same time and for distances up to about 0.1 m.

1 Relative displacement measurement

The position signal, generated in an all-optical fiber based sensor head that is placed

opposite to a reflector attached to the displacing target, is remotely collected by

means of an optical single mode fiber (SMF, figure 2). This offers the advantage of

non-invasive operation even in the harshest environments, as all electronic parts are

separated from the sensor head. The sensor head, having a size of only few

millimeters in order to fit into the tightest setups, is based on a patented confocal

technology that allows working ranges up to about 0.1 m with alignment tolerance of

0.4°, thus simplifying the optical alignment process. Furthermore, the simple

structure of the sensor head makes it robust against thermal drifts and easy to mount.

To recover both displacement and direction of the moving target with constant

sensitivity, the position signal is generated using a quadrature detection method based

Figure 1: Ultra-compact interferometric

displacement sensor with miniature sensor

heads (attocube FPS3010).

99



on the modulation of the laser wavelength. The modulation frequency is high enough

to track displacement velocities up to 1 m/s, as demonstrated in figure 3.

The interferometer is also suited for low and high frequency vibrometry applications,

as can be seen in figure 4 (the maximum detectable frequency can be further

increased using the FPS3010 high speed interfaces).

WR D1

D2

TunableLaser Cavity

Reflector

SMF2x2

Coupler

Figure 2: Schematic fiber-optic circuit diagram. The laser light (telecom wavelength)

is routed to the wavelength reference (WR) and to the interferometer axis via 2x2

directional couplers and is detected at detectors D1 and D2.

0 0.2 0.4 0.6 0.8 1-5

0

5

10

15

20

Time (s)

Dis

pla

cem

ent

(mm

)

0 0.2 0.4 0.6 0.8 1-0.5

0

0.5

1

1.5

Time (s)

Vel

oci

ty (

m/s

)

Figure 3: Displacement (left) and velocity (right) of a linear magnetic drive.

1 2 3 4 50

10

20

30

Frequency (Hz)

Am

pli

tude

(pm

)

4 pm

34 35 360

2

4

6

8

10

Frequency (kHz)

Am

pli

tude

(pm

) 8 pm

Figure 4: Digital displacement spectrum for a piezo modulated with 2.5 Hz (left, 0.8

kHz bandwidth) and 35 kHz (right, 100 kHz bandwidth). Data are recorded via USB

interface. Frequencies in the 1 MHz range are achievable with faster interfaces.

100


2 Distance range measurement

The measurement of absolute distances with nanometer repeatability is of special

interest for applications where the measurement is interrupted, causing a loss of

position. This can be due to a beam interruption, a power failure or intentional

interruption or simply when the application requires repeatable alignment of objects.

There are several successful approaches to this issue [2], but they suffer from system

complexity which again has a negative impact on the costs. The prototype presented

in the following sections makes use of the tunability of the laser currently used in the

FPS3010, thus allowing to extend the capabilities of the system beyond that of a

usual displacement interferometer towards an absolute distance interferometer

without the problem of position ambiguity (i.e. a laser range meter). In the following

sections, two different techniques compatible with the current interferometer system

FPS3010 are considered and evaluated.

2.1 Wavelength tuning

The simplest way to measure absolute distances with a laser interferometer is to scan

its wavelength. This induces a

phase change ΔΦ in the

interferometer cavity under test,

which is proportional to the cavity

length x, expressed by

n

cx

42, (1)

where c is the speed of light, n is

the refractive index of the medium

in the cavity and Δν is the change

of the laser emission frequency

(factor 2 because light traverses cavity twice). The main limiting error sources of this

technique are periodic nonlinearities due to the interferometer cavity, wavelength

uncertainties and phase noise. Their contribution to the total measurement uncertainty

are shown in figure 5. A further error arises when the cavity drifts during the

wavelength scan.

0 20 40 60 80 1000

10

20

30

40

Target distance x (mm)

Un

cert

ain

ty x

(

m)

Periodic non-lin.

Phase noise

Wavelength

Figure 5: Uncertainty of frequency tuning

interferometry.

101


2.2 Improved distance ranging

As the wavelength tuning technique suffers from a rather large uncertainty, we have

developed a proprietary distance ranging technology which we will publish later. Its

functionality is demonstrated in figure 6, showing the position reconstruction after a

system shut down for a constantly drifting target position. When measuring with a

displacement sensor, the position is zero after system reinitialization (left panel). In

our novel prototype the starting point is not lost as it can be seen in the right panel.

The prototype we built has presently an absolute position repeatability of ±2 µm, but

the next prototype will achieve a position repeatability in the nanometer range.

Acknowledgements

The authors acknowledge financial support from the Technische Universität

München (TUM) – Institute for Advanced Study (IAS), funded by the German

Excellence Initiative, and from the German Research Funding (DFG) through the

TUM - International Graduate School of Science and Engineering (IGSSE).

References:

[1] K. Karrai and P.-F. Braun. Multi-channel Optical Fiber Based Displacement

Metrology. Proceedings of the 11th euspen International Conference – Lake Como –

May 2011.

[2] J. R. Lawall. Fabry–Perot metrology for displacements up to 50 mm. J. Opt. Soc.

Am. A, Vol. 22, No. 12, December 2005.

0 10 20 300

1

2

3

4

Time (min)

Dis

pla

cem

ent

(

m)

0 10 20 300

2

4

6

8

Time (min)

Dis

pla

cem

ent

(

m)

Figure 6: Reconstruction of position after system shutdown. Left: Measurement of the

elongation of a drifting cavity with the FPS3010 displacement sensor. Right:

Reconstruction of the actual position with new prototype and corresponding error bar.

102


Design of a nanometer-accurate air bearing rotary stage for

the next generation nano-CT scanners

S. Cappa1, D. Reynaerts1, F. Al-Bender1 1 KU Leuven, Department of Mechanical Engineering, Belgium

[email protected]

Abstract

Micro- and nano-CT scanners are increasingly used in precision engineering. Proper

selection of the key components of these devices allows elimination of most artifacts

already during data acquisition. Typical nano-CT scanners contain an air bearing

rotary stage for object manipulation due to their rotational accuracy compared with

conventional bearings. In this work, the radial error motion of such an air bearing

rotary stage is reduced to 3 nm by optimising the feeding system, leading to a new

state of the art record for passive aerostatic bearings to our knowledge.

1 Introduction

Today, nano-CT scanners have a resolution as fine as 50 nm and are used in several

applications like life science studies, semiconductor industry and advanced material

analysis. They contain three main components: an X-ray source, an object

positioning system and an X-ray detector. However, the key parts are not ideal and

in most cases designed as a compromise between performance and price.

Performance limitations in the key components lead to artifacts in the acquired

angular projections and reconstructed slices. Most studies in the field have focussed

on increasing the accuracy of the X-ray source/detector and by compensating the

artifacts through acquisition and reconstruction software. However, there is still

much room for improvement by increasing the performance of the positioning

system. This is the objective of this work.

2 Air bearing design

The axis-of-rotation error motion of a well-designed aerostatic rotary stage is mainly

determined by the machining accuracy of the bearing surfaces as the clearances

103



should be as small as possible in order to obtain a high stiffness. The influence of

various manufacturing errors and bearing parameters on the radial error motion of an

aerostatic journal bearing is analysed in [1,2]. The results of this study show that the

running accuracy can be improved most effectively by increasing the number of

feedholes Nf of an air bearing system. As a result, the pressure profile between rotor

and stator is more uniform, which ultimately reduces the influence of irregularities

in the bearing surfaces.

However, the number of feedholes of an air bearing system is restricted from

practical point of view. A porous material, on the other hand, has an infinite number

of feedholes (ideally). As a result, a 2-DOF orbit model is developed to analyse the

radial error motion of a porous aerostatic journal bearing. The results were very

encouraging: 3 nm radial error motion.

3 Experimental validation

An existing air bearing rotary stage with inherent restrictors was adapted to a porous

type air bearing in order to validate the theoretical results. However, the thrust

bearings, each made up of eight inherent restrictors (Nf = 8), were not adapted. The

radial error motion is measured with the use of a novel reversal technique [3],

separating the artifact form error from the error motion. The measurement setup is

shown in Fig. 1.

Figure 1: Measurement setup.

104


The radial error motion of the rotary table, partially mounted in a granite block

which is placed on a high-precision indexation table, is measured at a rotational

speed of 60 rpm and a supply pressure of 5 bar. The rotor is manually driven as an

electrical drive system can have a significant influence on the performance. The

rotary table is equipped with a rotary encoder triggering the data sampling at evenly

spaced angular increments. In this way, the effect of spindle speed fluctuations is

eliminated.

The least squares synchronous radial error motion of the rotary table under test is

9 nm, as illustrated in the polar plot of Fig. 2. This result differs slightly from the

3 nm calculated by the orbit model. From Fig. 3 it can be seen that the harmonics at

n = k . Nf 1 with k N0 (grey) are remarkably higher than the remaining harmonic

components (black). This can be attributed to the tilt error originating from the thrust

bearings, which were not taken into account in the orbit model.

From the data in Fig. 3, it is apparent that the influence of the thrust bearings cannot

be ignored. As a result, a new rotary table, with the journal and thrust bearings each

made up of a porous feeding system, was designed and validated.

The synchronous and asynchronous radial error motion of this new design (full

porous) is compared with the first rotary table (journal: porous – thrust: restrictors)

for several supply pressures Ps in Fig. 4. It is apparent from this figure that both the

synchronous and asynchronous radial error motion is reduced to 3 nm and 2 nm,

respectively, by using a porous feeding system instead of inherent feedholes for the

Figure 2: Radial error motion

of the rotary table under test

(9 nm).

Figure 3: Frequency spectrum of the

synchronous radial error motion.

105


thrust bearings. These results nearly equal the noise level of the Elite Series

capacitive sensors of Lion precision used for the tests (1-2 nm).

4 Conclusions

In this work, the radial error motion of an aerostatic rotary table is reduced to the

noise level of the capacitive sensor, i.e. 3 nm. This is a new state of the art record for

passive aerostatic bearings to our knowledge. To achieve this, the supply geometry

of the air bearing rotary system was optimised by increasing the number of

feedholes by the use of a porous feeding system.

Acknowledgement:

This research is sponsored by the Fund for Scientific Research - Flanders (F.W.O.)

under Grant G037912N. The scientific responsibility is assumed by its authors.

References:

[1] Cappa, S., Waumans, T., Reynaerts, D., Al-Bender, F. Theoretical study on the

radial error motion of high-precision aerostatic rotary tables. Proceedings of the 11th

euspen International Conference (pp. 307-310).

[2] Cappa, S., Waumans, T., Reynaerts, D., Al-Bender, F. Reducing the error motion

of an aerostatic journal bearing. Proceedings of the 12th euspen International

Conference (pp. 435-438).

[3] Cappa, S., Reynaerts, D., Al-Bender, F. (2012). A new Spindle Error Motion

Separation Technique with sub-nanometre uncertainty. Proceedings of the 12th

euspen International Conference (pp. 141-144).

Figure 4: Comparison of the synchronous and asynchronous radial error motion

of a full porous and partially porous aerostatic rotary table.

106


Practical method for determining the metrological structure

resolution of dimensional CT

S. Carmignato1, P. Rampazzo1, M. Balcon2, M. Parisatto3 1 University of Padova, Department of Management and Engineering, Italy 2 University of Padova, Department of Industrial Engineering, Italy 3 University of Padova, Department of Geosciences, Italy

[email protected]

Abstract

This work deals with a practical approach for determining the metrological structure

resolution in X-ray Computed Tomography (CT) for dimensional measurements.

Advantages over other applicable approaches are discussed. The experimental results

obtained from the implementation of the method using a micro-CT system are

compared with the geometrical unsharpness of CT reconstructions.

1 Introduction

In CT dimensional metrology, the metrological structure resolution describes the size

of the smallest structure that can still be measured without exceeding specified error

limits [1]. The metrological structure resolution should always be tested and specified

in addition to other relevant metrological characteristics, such as length measurement

errors and probing errors. It provides additional important information: for example,

if smoothing filters are increased, the probing error of form can be improved while at

the same time the structure resolution being worsened [2].

Spatial resolution in computed tomography has already been studied thoroughly in

literature and several methods for its determination have been published and

standardized [2]. However, these methods refer to spatial resolution in the grey-scale

voxels (volumetric pixels), which does not take into account the complete

measurement chain of CT dimensional measurements, while the metrological

structure resolution does so (e.g. it takes into account also threshold determination,

surface points extraction and filtering and averaging of surface points). Achieving a

good spatial resolution in the grey-scale voxels is necessary but not sufficient for

achieving also a good metrological structure resolution.

107



New practical methods are demanded in industrial CT metrology for fast and reliable

determination of metrological structure resolution, encompassing the whole

measurement chain [3].

2 Approach for determining the metrological structure resolution

The proposed approach is based on the ‘Hourglass’ reference standard: two calibrated

spheres with the same nominal diameter (D), enclosed in a carbon fibre tube and

physically touching each other as shown in Figure 1-a. The geometry of the sample

was chosen as simple as possible, in order to facilitate the evaluation of the structure

resolution. Due to the finite structure resolution, the dimensions d and h of the

distorted contact zone in the CT surface reconstruction (see Figure 1-c) increase as

the structure resolution increases (see Figure 1-b).

The ‘Hourglass’ standard was preliminary introduced in [2]. A similar concept, using

a microtetrahedron sample, was proposed also by Bartscher and Härtig [4] to the ISO

TC 213 WG10, but not yet accepted for adoption in the ISO standard on verification

of CT measuring systems, which is currently under development.

Using the ‘Hourglass’, the structure resolution is determined by measuring the height

h on the surface reconstruction. For better accuracy, the value of h can be calculated

indirectly from the values of the diameters D and d, since these diameters can be

measured with lower relative uncertainty.

Figure 1: (a) X-ray image of the ‘Hourglass’ standard, with spheres having diameter

D = 8 mm. (b) Three surface reconstructions obtained from three different CT

measurements of the ‘Hourglass’, with increasing structure resolution from left to

right. (c) Schematic representation of diameter d and height h of the contact zone of

the surface reconstruction resulting from CT measurement of the ‘Hourglass’.

108


3 Comparison with geometrical unsharpness

An experimental investigation was conducted using a micro-CT system SkyScan

1172 (SkyScan-Bruker microCT, Kontich, Belgium). The ‘Hourglass’ standard was

measured using the micro-CT system with five different magnifications, obtaining

five CT measurements with different voxel sizes: 5.6 µm, 12.5 µm, 22 µm, 27 µm

and 50 µm. Each measurement test was repeated three times, for a total of 15

measurement tests performed on the ‘Hourglass’.

For each CT measurement test, the surface geometry of the ‘Hourglass’ was

reconstructed and a point cloud was extracted. Each point cloud was analysed using

specific elaboration procedures implemented in three-dimensional data modelling and

evaluation software (PolyWorks, InnovMetric Software Inc., Canada), determining

the actual diameter d and height h of the contact zone (Figure 1-c).

The values of the height h were compared to the values of the geometrical

unsharpness of the respective CT scans. A simplified estimation of the geometrical

unsharpness was used [5]:

√

(1)

where UF is the unsharpness caused by the finite focus size (SF), and UD is the

unsharpness caused by the finite pixel size of the detector (SD). They were assumed

respectively equal to: – and , where m is the

geometrical magnification, defined as the ratio between the source-to-detector

distance and the source-to-rotation-centre distance.

For each CT measurement test performed, the ratio between the height h measured on

the ‘Hourglass’ and the geometrical unsharpness estimated according to equation (1)

was computed. For the 15 measurement tests, the mean value of this ratio was found

equal to 1.1, with standard deviation equal to 0.15. The variability of this result was

due mainly to the influence of ring artifacts, disturbing the evaluation of the height h

and the diameter d of the contact zone of the surface reconstruction resulting from the

CT scan of the ‘Hourglass’.

4 Discussion and conclusions

The proposed method using the ‘Hourglass’ reference standard is definitely more

easily applicable than the basic method currently proposed in the guideline VDI/VDE

109


2617 Part 13 [1], which consists in determining the diameter of the smallest sphere

for which the measuring system is able to determine a diameter, with error stated by

the CT system manufacturer. The latter method, in fact, may need a large number of

calibrated microspheres with different diameters, which definitely are more difficult

to be calibrated, handled and CT measured than the two spheres used in the

‘Hourglass’.

The experimental investigation carried out in this work demonstrated that the

‘Hourglass’ approach can be used effectively and efficiently for determining the

metrological structure resolution. The investigation showed also that, for the specific

conditions used in this work, the ratio between the height h of the contact zone

measured on the ‘Hourglass’ and the estimated geometrical unsharpness of the CT

scan is equal to 1.1 on average. Future work should include the following studies:

relation to the spatial frequency response of the instrument, investigations using

different CT systems and scanning parameters, investigations using samples with

different materials and dimensions, and investigations on the influence of different

sample orientations and the influence of ring artifacts.

References:

[1] VDI/VDE 2617 - Part 13 (2011). Guideline for the application of DIN EN ISO

10360 for CMMs with CT-sensors. VDI, Duesseldorf.

[2] S. Carmignato, A. Pierobon, P. Rampazzo, M. Parisatto, E. Savio, (2012). CT

for industrial metrology - Accuracy and structural resolution of CT dimensional

measurements. Proc. of the Conference on Industrial Computed Tomography

(ICT); pp. 161–172.

[3] M. Bartscher, H. Bremer, T. Birth, A. Staude, K. Ehrig, (2012). The resolution

of dimensional CT - an edge-based analysis. Proc. of the Conference on

Industrial Computed Tomography (ICT); pp. 191–200.

[4] M. Bartscher, F. Härtig, (2011). ISO TC 213 WG 10 study on the structural

resolution of CT systems for dimensional measurements. ISO TC 213 WG10

meeting; Bejing, September 2011.

[5] P. Müller, J. Hiller, A. Cantatore, M. Bartscher, L. De Chiffre, (2012).

Investigation on the influence of image quality in X-ray CT metrology. Proc. of

the Conference on Industrial Computed Tomography (ICT); pp. 229–238.

110


Traceable profilometer with a piezoresistive cantilever for

high-aspect-ratio microstructure metrology

M. Xu, U. Brand, J. Kirchhoff

Physikalisch-Technische Bundesanstalt (PTB), Bundesallee 100, 38116

Braunschweig, Germany

[email protected]

Abstract

A profilometer is developed to traceably characterize the roughness of high-aspect-

ratio micro structures. A slender silicon cantilever with an integrated piezoresistive

strain gauge is used in the instrument for sensing the surface and signal read-out.

With a width down to 30 µm and a length up to 5 mm, the cantilevers allow to

measure the roughness profiles inside of micro holes with diameters of 100 µm and

less. At the head of the profilometer three laser interferometers with 1 nm resolution

are arranged perpendicularly to each other to provide metrological traceability. For

step height and roughness arithmetical mean deviation measurements, the uncertainty

of the system is within ±10 nm. Finally, the roughness profile inside a micro hole of

100 µm in diameter is successfully measured.

1 Introduction

High-aspect-ratio microstructures (HARMS), such as micro holes, micro pipes and

micro gears, are used in practice in fields of biotechnology, aerospace and

automotive industries. However, quality control of HARMS, especially traceable

metrology at the nanoscale is still a challenge because it is difficult to measure such

structures by using existing measurement technology. The high walls with extremely

narrow spacing or the inside profiles of deep micro holes are neither detectable with

tactile nor with optical methods. In many cases these structures are only measured

after sectioning of the parts. Therefore a traceable profilometer has been developed

to achieve a non-destructive and accurate inspection of HARMS. In this paper the

construction of the profilometer is at first introduced, then the measurement of

roughness profiles inside of a micro hole with 100 µm diameter is performed.

111



2 System design and construction

Key part of the profilometer is a long silicon cantilever with an integrated

piezoresistive strain gauge (Figure 1), which is developed and is described elsewhere

[1, 2, 3]. The slender cantilever is used for signal read-out and enables the ability for

high-aspect-ratio surface measurement. The cantilevers are 1.5, 3 and 5 mm long, 30,

100 and 200 µm wide. The tips are about 25 µm, 25 µm and 70 µm high. Tip radii are

smaller than 0.1 µm. The sensitivity of the cantilevers are about 0.31, 0.19 and 0.24

µV/nm, and the noise of the cantilever read out is about 3 nm in a bandwidth of 10

mHz to 1 kHz without acoustic noise isolation in a laboratory room.

The actual construction of the profilometer is given in Figure 2. The cantilever is

bonded on a cantilever holder and mounted on the XYZ piezo stages with a motion

range of 800 µm × 800 µm × 250 µm (x × y × z). The cantilever moves with the

piezo stages. The read-out of the cantilever serves as the control value after

amplification and is input to the z piezo stage controller. During measurement the

contact force between cantilever and artefact surface keeps constant and can be set

down to 1 µN.

The artefact is placed on a coarse positioning stage with a movement range of 50

mm × 50 mm × 12 mm (x × y × z). The three linear coarse stages are mounted on a

rotation stage so that the artefact can be rotated by 360 degrees around the z-axis.

All coarse stages are equipped with motorized drives. In addition, the entire head of

the system can be moved about 100 mm in the z-direction to measure large artefacts

up to 8 cm × 10 cm × 10 cm.

On the head of the system, three laser interferometers with 1 nm resolution (SIOS,

SP2000) are arranged perpendicular to each other to provide metrological traceability.

The three laser beams intersect on the cantilever tip to achieve an Abbe error-free

measurement. The measured artefact surface is constructed by the readings of the

three laser interferometers.

Through numerous comparison measurements of step heights and roughness

standards with other metrology instruments in PTB, it is found that the error of the

112


profilometer for step height and roughness arithmetical mean deviation Ra

measurements is within ±10 nm (2σ).

Fig. 1 A 1.5 mm long piezoresistive

cantilever

Fig. 2 The profilometer with

piezoresistiver cantilever

roughness sensor

(a) (b)

Fig. 3 The measurement of a micro hole by the profilometer: (a) the measurement

artefact, a steel plate with micro holes of 100 µm diameter and (b) the 1.5 mm long

piezoresistive cantilever in a micro hole during a measurement.

3 Roughness measurements inside a micro hole

Micro holes with 100 µm diameter each were drilled into a steel plate arranged in a

5x3 matrix. The depth of the holes was 500 µm (see Fig. 3a). The roughness profile

inside one of the micro holes was measured using a 1.5 mm long piezoresistive

cantilever. During the measurements the cantilever moves along the axis of the

113


cylindrical micro hole with a speed of 20 µm/s. The probing force of the cantilever

during the measurement was 7 µN.

A 340 µm long profile was measured with 3580 sampling points and the

measurement was repeated eight times. Two repeated measurements are shown in

Fig. 4. Using a short wavelength noise filter λs = 2.5 µm, the measured profiles were

evaluated with the Reference Software RPTB for roughness measurements. An

arithmetic mean roughness of Ra = 845 nm at a standard deviation of = 2.0 nm

was obtained.

Fig. 4 A typical roughness profile measured inside a micro hole of 100 µm diameter

References:

[1] Peiner E, Balke M, Doering L and Brand U 2008 Tactile probes for dimensional

metrology with microcomponents at nanometer reolution Meas. Sci. Technol. 19

064001

[2] Peiner E, Balke M, Doering L, Brand U, Bartuch H and Völlmeke S 2007

Fabrication and test of piezoresistive cantilever sensors for high-aspect-ratio

micro metrology Proc. Mikrosystemtechnik Kongress 2007 (Dresden, Germany,

15–17 October) (Berlin: VDE) 369-74

[3] Peiner E, Balke M and Doering L 2008 Slender tactile sensor for contour and

roughness measurements within deep and narrow holes IEEE Sensors Journal 8

1960-7

114


Verification of thickness and surface roughness of a thin

film transparent coating

K. Mohaghegh1, H.N. Hansen

1, H. Pranov

2, G. Kofod

2

1Technical University of Denmark, Denmark 2InMold Biosystems, Denmark

[email protected]

Abstract

Thin film coatings are extremely interesting for industries, where there is a need to

protect a highly accurate surface which has tight dimensional tolerances. The topic is

important both in the production of new metallic tools and repair applications. In both

applications it is vital to have a specific knowledge about coating thickness and

roughness. In the present paper a novel application of a transparent HSQ coating is

presented. Furthermore the thickness and roughness of the transparent coating with

nominal thickness of 1 µm is measured in the presence of surface roughness of the

substrate. Measurements were done using AFM and a precision 3D mechanical stylus

instrument.

1 Introduction

Polishing of metal tools and parts is a manual process with many drawbacks and

risks, including being detrimental to worker health. Therefore it is relevant to

investigate alternative methods for polishing, such as methods relying on coatings.

Thin film Hydrogen Silsesquioxane (HSQ) is commonly used in the semiconductor

industry in the manufacture of integrated circuits (ICs), both as a low-k dielectric and

as a planarization material to fill in the gaps between metal wires and spatially

separated components [1]. HSQ is an unstable, cage-like silane hydrolyzate, which

cures to form a solid, amorphous quartz layer. It can be obtained as a pure material or

in liquid form prepared for IC manufacture, which can be applied via typical means

of coating, leading normally to reductions in the surface roughness [2]. The focus of

this paper is the geometric characterisation of the thin HSQ coating applied on a flat

surface of steel.

115



2 Measurement of coating thickness

A flat surface of a gage block made of steel has been used as the base surface for

application of the HSQ coating. The HSQ coating was applied using an ultrasonic

spray nozzle, and then it was cured at 450°C for 1 hour. For the sake of height

comparison after coating, a portion of the surface was masked before coating. So a

step was created which was the subject of height measurement. The instrument used

was Form TalySurf 50 inductive stylus profiler with 0.6 nm height resolution, 250 nm

lateral resolution and 2 µm tip radius capable of 3D movement. The stylus covered a

length of 1.5 mm with a width of 50 µm. Figure 1 shows the result of measurement

after applying a Gaussian filter (λs = 2,5 µm) and plane correction in software SPIP,

version 6.0.13 (Image Metrology A/S, Horsholm, Denmark).

Step Height

Figure 1: Height measurement of HSQ layer by 3D stylus including the height

histogram

An area in the middle of the step (about 200 µm length in each side) shows some

shape irregularities which are mainly due to the reflow phenomenon created on the

step. But the remaining area which is sufficient for measurement (about 500 µm in

each side) covered in this test shows a quite homogeneous height distribution of 0.6 ±

116


0.07 µm with respect to base surface (left distribution in the histogram) which itself

exhibits ±0.12 µm height deviation. An improvement in surface roughness after

coating is demonstrated by comparison of different distributions although mechanical

stylus might not be the proper instrument in this range of roughness.

An additional effort to support coating thickness measurement has been performed to

get a cross sectional information about the coating layer through Jeol 5900 scanning

electron microscope (Fig.2). In order to recognize the HSQ layer, a thin layer of

Nickel (400 nm) was deposited onto surface of the work piece using Physical vapour

deposition (Metallux-ML18). Compared to the large coverage of the stylus (1500

µm), the 1:1 SEM picture covers only a very small portion on the surface (40 µm).

The coating thickness variation in the presence of very high lateral resolution of SEM

gives a very different distribution although the nominal value is still in conformity to

what is measured by mechanical stylus.

Figure 2: SEM image of the substrate (bottom) coated with HSQ (middle black layer)

and an additional Ni layer (white thin layer in the middle of the picture)

3 Measurement of roughness

Roughness tests were done using atomic force microscope [3]. Scanning areas of

50X50 µm were selected on both uncoated and coated surfaces (Fig.3). The 3D

visualisation of the surfaces shows a considerable improvement in surface roughness

after coating (Fig.3A). The nature of the coated surface is completely different to the

initially polished one. The 3D roughness parameters as well as the histogram (Fig.3B)

demonstrate this change (Sa reduced from 13.7 to 4.7 nm and Sz reduced from 183.7

to 56.2 nm). The bearing curve parameters clearly show an improvement on the

surface in a lower height distribution. The specific values of peak, valley and core

117


roughness parameters are all reduced simultaneously (Fig.3C) meaning that the

reduction of height is distributed evenly between peaks, core and valleys.

A parallel roughness study has been performed by a 2D mechanical stylus (2 µm tip

radius) with 12 repetitions covering an evaluation length of 1.25 mm. After applying

Gaussian filters (λs = 2,5 µm, λc = 250 µm) it resulted in 10.2 ± 0.4 nm Ra for the

uncoated surface and 8.75 ± 0.4 nm Ra for the coated surface. An examination of the

background noise of the stylus on a reference plain glass showed 4 nm Ra which

makes it difficult to rely on stylus results on this range of roughness.

B

C

A

B

C

Raw surface Coated surfaceSa = 4.7 nm

Sz = 56.7 nm

Sa = 13.7 nm

Sz = 183.7 nm

Figure 3: 3D surface characterisation of raw (left) and HSQ coated surface (right):

visualisation (A), height histogram (B) and bearing curves (C)

4 Conclusion

The study toward surface characterisation of the HSQ coating in the presence of a

base roughness shows that the thickness of the coating is 0.6 ± 0.07 µm based on long

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coverage of 3D stylus (1500 µm), while short coverage surface variation (50 µm) at

each side of the step as measured by AFM gives 183 nm Rz on the uncoated surface

and 56 nm Rz on the coated surface. The smoothening effect of the coating is

obvious through the study which is the subject of an upcoming publication by the

authors. SEM microgarphs compared to AFM and stylus instruments exhibit a

different view of thickness variation because of the high lateral resolution. The

advantage of SEM is the possibility to observe inner and outer coating surfaces at

exactly the same point on the surface which is quite valuable to study the correlation

of coating roughness based on a certain substrate roughness.

Acknowledgment

The paper reports the work undertaken in the context of the project “dimensionally

stable reflow polishing of molding tools” (RePol) which is founded by The Danish

National Advanced Technology Foundation.

References

[1] P.S. Ho, J. Leu, W.W. Lee, Low Dielectric Constant Materials for IC

Applications, Springer, Berlin, 2003

[2] H. Pranov, Reactive silicon oxide precursor facilitated anti-corrosion treatment,

WIPO Patent WO/2013/017132, 2013

[3] DME DualScope 95, Danish Micro Engineering A/S, Denmark, 2007

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Measurement and Evaluation Processes for Inner Micro

Structures

T. Krah1, A. Wedmann1, K. Kniel1, F. Härtig1 1Physikalisch-Technische Bundesanstalt, Braunschweig und Berlin, Germany

[email protected]

Abstract

Inner micro structures can be found in a growing number of products, especially in

medical equipment. Frequently, the quality requirements are very high, whereas the

metrological possibilities are limited. The new T-shaped micro probe presented in

this article forms an approach to solve this discrepancy. It allows high accurate tactile

measurements at internal micro structures such as inner threads. First verification

measurements are performed with the new probe applied in a 3D coordinate

measuring machine (CMM). Furthermore, an approach for a laminar evaluation of the

measured thread flanks is presented.

1 Introduction

In economic terms, a constant growth of the market for technical and medically

functional components from the micro systems technology sector can be observed.

These components are moulded by complex structures, whose accessibility frequently

poses problems, especially when it comes to the employment of measurement

devices. It is e. g. not uncommon for inner micro threads that inner micro structures

with a dimension of less than 0.2 mm can be found. However, calibrations traceable

to the SI units can only be performed until a minimal thread size of M3, the main

reason being the lack of micro probing processes to facilitate a way of probing the

complex inner structures. To eliminate this problem, a complete process chain for the

calibration of complex inner micro structures has been developed by the

Physikalisch-Technische Bundesanstalt (PTB), comprising the development and the

implementation of innovative, robust T-shaped micro probes, their adaptation into a

coordinate measuring machine (CMM) by means of especially designed calibration

processes, and the performance of laminar analyses.

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2 Design and fabrication of new T-shaped micro probes

The innovative T-shaped micro probes are characterised by the fact that the probe

spheres are basically clamped and not, as hitherto common, glued or brazed to the

shaft. By using this new method, not only a high stability, but also the possibility to

replace spheres once they are worn out, and to reuse the shaft can be obtained. The

original design of the micro probe is based on the patent [1] (Figure 1). The ideal

stylus designs for different probe sphere diameters were identified with the help of

FEM analyses, taking into consideration the stylus length, the elasticity, and the

processibility. The objectives behind the design optimization were a maximization of

the clamping force on the probe sphere and the avoidance of stress peaks.

Figure 1: Design of a T-shaped micro probe (photograph and corresponding sketch)

Different probes were manufactured via micro wire cut EDM and die sinking. Special

assembly jigs for the insertion of the probe spheres were developed and produced.

Based upon this functional principle, it is currently possible to implement T-shaped

micro probes with a probe sphere diameter of down to 120 µm. The geometric

dimensions can be varied in a wide range. For probes with a probe sphere diameter of

120 µm the length of the shaft is set to 1.82 mm, its cross-section to 200 x 250 µm

and the stylus constant (length between the spheres’ outside [2]) to 490 µm.

Due to its design the T-shaped micro probe shows a highly anisotropic probing

behavior. In order to get proper measurement results it is important to have a good

knowledge of the probe’s mechanical behavior. A universal characterization method

for 3D tactile probing systems that is also suited for microprobes is described in [3].

3 Measurement and evaluation of inner micro structures

3.1 Measurement setup

The new T-shaped micro probe was adapted to a standard CMM. System parameters

such as probing force and dynamics were adjusted to get reliable measurement results

and simultaneously not do damage the probe. First verification measurements were

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carried out on a calibrated thread measurement standard with the dimension M3 x 0.5,

which barely can be calibrated with common probes. Further measurements were

carried out with so far uncalibrated micro threads with the dimensions M10 x 0.175,

M0.9 x 0.175, and M0.7 x 0.175.

3.2 Standard conform evaluation

The measurements of the micro threads were carried out in the same way as

measurements of macroscopic thread measurement standards [2]. The diameters were

calculated by subtracting the x- respectively y-components of three measured points

on opposite sides of the thread. For the calculation of the pitch diameter it is

important that the probe sphere touches the thread on both sides of the gaps. Special

attention was paid to the touch behavior of the micro probe in gaps. In microscopic

dimensions parameters such as the roughness of surfaces, particles and fluidic layers

have a much stronger impact on getting in contact on both sides of the thread gap

than in macroscopic dimensions. The visual observation and the analysis of repeated

measurements showed that a reliable contact of both flanks in the gap was performed.

The measurement results of the micro threads are in excellent accordance with the

calibration results within the respective measurement uncertainties. Harmonious

results were also achieved with first measurements of micro threads of the sizes M10,

M0.9 and M0.7, each with a pitch of 0.175 mm, after applying the standard conform

measurement strategy and evaluation of the customary thread parameters pitch, pitch

diameter as well as inner or outer diameter.

3.3 Laminar evaluation

The laminar analysis of flank areas by dint of least squares algorithms may form the

basis to a future evaluation of thread geometry in its entirety. For this purpose,

measurement points spread all over the thread flank were recorded with an increased

density. By applying an innovative laminar evaluation process for the first time a

complete 3D analysis of the functional areas can be carried out. Based on this, it is

now possible to make statements with regard to factors such as periodic pitch errors,

convexities of the flank areas or local defects.

The laminar evaluation is executed in three steps. First, the flank areas are unwinded.

Since threads have a linear profile and pitch the unwinded surface is a plane.

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Depending on the used measuring device and yielded presentation of the coordinate

points an additional coordinate transformation may become necessary. In a second

step, a plane is fitted into the measured points. In a last step, all desired parameters

are calculated and compared to nominal values. Figure 2 shows the example of a

laminar evaluation for a plug thread gauge of the size M64 x 6. From this figure it is

possible to extract a clear periodic pitch error, a slight error of the flank angle and a

very small pitch error, whereas formerly through the application of a standard

conform evaluation to the measured plug only the pitch error could be recognized.

Figure 2: Laminar evaluation of threads using the example of a plug thread gauge of

the size M64 x 6

Acknowledgement

The presented work was financed by the Federal Ministry of Economics and

Technology within the MNPQ project “Rückführbare und robuste Kalibrierverfahren

für Mikroinnenstrukturen”. The authors would also like to thank the project partners

Co. Emuge, Co. Decom and Co. Lehren- und Messgerätewerk Schmalkalden for their

supply with probes and several test specimens.

References:

[1] Patent: DE10 2011 050 257 A1 2012.11.15 Tasteinrichtung zum Antasten von

Oberflächen sowie Verfahren zur Herstellung einer solchen Tasteinrichtung.

[2] EURAMET cg-10, Version 2.0 (2011): Determination of Pitch Diameter of

Parallel Thread Gauges by Mechanical Probing.

[3] N. Ferreira, T. Krah, K. Kniel, S. Büttgenbach, F. Härtig: Universal

characterization method for 3D tactile probing systems. ISPEMII 2012, China, 2012

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Quantitative assessment of nano wear of DLC coated samples

using AFM and optical confocal microscopy

G. Dai1, F. Pohlenz1, H. Bosse1, A. Kovalev2, D. Spaltmann2, M. Woydt2 1Physikalisch-Technische Bundesanstalt (PTB), Bundesallee 100, D-38116

Braunschweig, Germany 2 Federal Institute for Materials Research and Testing (BAM), D - 12203 Berlin,

Germany

[email protected]

Abstract

Nano wear of high-performance DLC-based thin film coatings, which had undergone

tribological tests under rolling/slip-rolling conditions, was quantitatively assessed

using calibrated AFM and confocal microscopy. Different areal S-parameters such as

arithmetic average roughness Sa, root mean square roughness Sq, surface skewness

Ssk, as well as areal V-parameters (related to Abbott-Firestone curve) have been

evaluated from the AFM images measured from the worn and unworn areas. The

cumulative distributions of surface area, projected area and material volume (loss)

were analyzed in addition. The study shows that for the analyzed S-parameters the

skewness Ssk was the most sensitive and reliable parameter to indicate the very small

wear-related changes of the DLC-coated surface.

1 Introduction

The amorphous diamond-like carbon (DLC) coating is well known for its high

hardness and wear resistance as well as low friction coefficients. It has been

increasingly applied, in particular, in the automotive industry to increase the lifetime

of car components and to reduce fuel consumption. Coating of rolling/slip-rolling

parts is subjected to cyclic fatigue and furthermore under deficient lubricant is rarely

reported and remains nowadays an attractive challenge [1].

Until today there is no single approach which can provide a complete and simple

description of the surface topography to reveal the fine changes on its asperity peaks

due to friction. In particular, it remains a challenging task for so-called “zero-wear”

processes, where the material loss due to wear is within the height range of the

original topography [2]. Consequently, a reliable quantitative analysis of such wear

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characteristics becomes crucially important to understand the “zero-wear” conditions,

which may significantly increase the wear resistance and longevity of friction

components. Quantitative assessment of nano wear of a DLC coating, which had

undergone rolling/slip-rolling wear testing, is studied in this paper.

2 Workpieces used for wear tests

The substrates were made of the quenched and tempered steels 100Cr6H (OVAKO

„PBQ‟), whose hardness is in the range of 66 HRC (Rockwell hardness C). Such

steels are typically used in bearings and serve as a reference for the investigations of

DLC-coated and uncoated novel steel grades. The DLC coating used for this study

was deposited using a pulsed vacuum arc deposition system. The steel substrates were

treated by an ultrasonic cleaning in an alkali water solution. The sample was sputter-

cleaned prior to deposition. The thickness of DLC coating was 2-3 µm, coating type

is a-C:H. The operating parameters of the friction test were as follows; slip-rolling

scheme with a difference of rolling velocity of 10%, the average pressure was 1.5

GPa, the number of revolutions over the duration of the test were 10 million cycles,

the lubricant VP1 SAE 0W-40 was applied to the contact zone.

3 Surface characterisation

An optical confocal laser scanning microscope (“LEXT OLS 4000” of the company

Olympus) with a short light wavelength of 405 nm was applied for fast overview

measurements, and an atomic force microscope (AFM) (“Dimension Icon” of the

company Bruker) was employed for detailed quantitative measurements at the local

areas of interests in this study. For both instruments, the amplification and linearity

of the scales was traceably calibrated by applying a set of step heights and lateral

gratings calibrated by a metrological AFM. Fig.1 presents two AFM images of the

DLC surface at the worn and unworn areas. The general surface structure

(morphology) seems to be similar, but the surface of the worn area (Fig. 1b) has

some flat areas (marked with circles) on the top of asperities that are not present on

the unworn surface shown in Fig. 1a.

These flat areas are the result of friction on the top of asperities without significant

destruction of the latter. It should be noted here that most of asperities have not

undergone such changes on their asperity tips during friction.

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Figure 1. AFM images of unworn (a) and worn surface (b) of the tested DLC

coating. Some flat areas on the top of asperities of the worn surface are marked.

4 Analysis of AFM measurement data

Different areal S-parameters such as arithmetic average roughness Sa, root mean

square roughness Sq and surface skewness Ssk were evaluated from the AFM images

measured on the worn and unworn areas [3]. In order to reduce the influence of the

surface waviness on the evaluation, a 2D Gaussian filter with a cutoff wavelength of

about 1.25 µm along the x and y axes was applied to pre-process the raw AFM

images. Surface parameters characterised at 21 different measurement locations at

the worn and unworn areas were compared. The result suggests that the skewness

Ssk was the most sensitive and reliable parameter for assessing the nano wear. As

shown in Fig. 2, the Ssk value is clearly reduced from the unworn areas to the worn

areas, which indicates the loss of peak structures of the asperities during the friction

test, agreeing well with the physical understanding of the “zero-wear” process. The

reliability of the measurements has been investigated and confirmed. The standard

deviation of the Ssk values of 8 repeat measurements on the same area is only

0.0003, much smaller than the wear induced Ssk change, 0.49. Moreover, the

relationship between Ssk and the material volume loss is simulated by a wear model

which assumes the removal of the top of asperities of worn surface during the wear

test, based on which the volume loss of about 2.5 x 107 nm3 per measured surface

area of 5 x 5 µm2 is estimated.

However, areal S-parameters mentioned above cannot directly reveal the tribological

behaviour and wear phenomena of the surface, which should be better interpreted by

surface features such as local contact spots, the distribution of real areas of contact

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between the rough surfaces, mean curvature radius of asperities, and root-mean-

square (Rq) of asperity peaks. To overcome this limitation, other surface

characteristics such as bearing projected area, bearing surface area and bearing

material volume were proposed and investigated. The bearing projected area curve is

the cumulative distribution of areas of material, intersected by a virtual cutting plane

along the height range of surface structures. It has a strong relation to the important

functional property of the surface such as the real contact area. In addition, the

bearing surface area curve is calculated as the sum of surface areas which exceed the

virtual cutting planes. The curve is valuable in tribology for calculating the adhesion

interaction of rough surfaces. The bearing material volume curve is calculated as the

sum of material volumes which exceed the virtual cutting planes. The curve is

crucial for estimating the volume loss of surface due to wear. As an example, the

bearing material volume curves calculated from a worn and unworn area are shown

in Fig. 3. Their difference especially in shape at the “peak zone” is clearly visible.

Further studies will be carried out to quantitatively assess nano wear from these curves.

The work was performed within the joint research project “JRP MADES” funded by

the European Metrology Research Programme (EMRP).

References:

[1] C.-A. Manier et al. Wear, vol. 268, no. 11–12, pp. 1442–1454, May 2010

[2] P. Pawlus and J. Michalski, Wear, vol. 266, no. 1–2, pp. 208–213, January. 2009

[3] ISO 25178-2:2012, Geometrical product specifications (GPS) -- Surface texture:

Areal -- Part 2: Terms, definitions and surface texture parameters

Figure 2. Characterised surface parameter

Ssk calculated at 21 different locations

selected at the worn and unworn

surface

Figure 3. Bearing material volume

curves evaluated from the worn and

unworn surface area of of 5 x 5 µm2

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Measurement Setup for the Experimental Lifetime

Evaluation of Micro Gears

G. Lanza1, B. Haefner1

1wbk Institute of Production Science, Karlsruhe Institute of Technology (KIT),Germany

[email protected]

Abstract

Micro gears are crucial parts of micro transmissions for various applications in

industries such as medical, automotive and industrial automation that require highest

precision. In order to enhance the lifetime prediction of micro gears, an experimental

approach is to be developed at the Karlsruhe Institute of Technology to model the

influence of geometric shape deviations and the material structure of micro gears on

their lifetime. For this, a highly precise experimental setup is required to conduct

abrasive experiments under clearly defined conditions. In this article a suitable

experimental rig is presented.

1 Introduction

Nowadays, micro transmissions are used in combination with micro motors in

manifold industrial applications such as dental drills or hexapod micro positioning

systems for wafer processing. Micro gears, defined as gears with a module < 200 µm

[1], are parts of micro transmissions, in which the gear quality is critical to the

functionality of the transmission. To ensure proper operation of the micro gears, a

reliable prediction of their lifetime is crucial. Lifetime evaluation is particularly

important for micro gears, as the influence of their geometric shape deviations on

their load-carrying capacity is significantly higher in comparison to gears with larger

modules. This is a consequence of their manufacturing processes, which are not

capable of producing micro gears with the same relative accuracy as larger gears.

The lifetime of micro gears is to be evaluated by an experimental approach at the

Karlsruhe Institute of Technology. After geometric measurements, a pair of micro

gears is systematically worn under realistic, clearly defined conditions, until a defect

of one of the micro gears can be detected. For the lifetime experiments and deduction

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of the lifetime model, a highly precise measurement setup is required to wear micro

gears up to an estimated torque of 1 Nm and a rotational speed of 3000 rpm. In

particular, accurate positioning of the micro gears to each other and stable locations

of their rotary axes have to be guaranteed within the range of few micro meters.

2 Literature review

A literature review shows that various rigs have been developed for different gear

experiments. However, only very few deal with gears with very small modules. Beier

developed a rig for lifetime experiments of assembled planetary gear transmissions

consisting of planetary gears with a module of 400 µm [2]. Braykoff designed a test

rig for gears with a module down to 300 µm to analyze their load-carrying capacity

[3]. However, only the experimental setup developed by Hauser is applicable for

micro gears, which was demonstrated for a module of 169 µm [4,5]. Precise

positioning of the micro gears is realized by a 5-axes manipulator and air bearings.

However, as the rig was designed for single-flank working tests, it is only applicable

for a low torque load of < 50 mNm, which is not adequate for the desired lifetime

experiments.

3 Measurement setup for the lifetime evaluation of micro gears

The developed measurement setup shown in figure 1 provides the functionality to

systematically abrade various types of pairs of micro gears under clearly defined,

variable conditions. The rotational speed, the torque as well as the center distance of

the micro gears can be precisely adjusted. Hence, the measurement setup is both

accurate and flexible with regard to different kinds of micro gears. It, however, is

restricted to cylindrical gears, which are by far the most common type of micro gears.

Variable rotational speed is provided by a synchronous motor (up to 5000 rpm), while

a hysteresis brake is used to adjust the torque (up to 3 Nm). Both rotational speed and

torque are measured by means of a sensor (relative uncertainty < 1 % for speed,

< 0.1 % for torque). Feedback control is implemented to guarantee constant

experimental conditions of the speed and the torque.

The bearings of the brake and the sensor are joined by a coupling (cf. figure 1). The

sensor as well as the motor is connected to a shaft. The micro gears to be used in the

lifetime experiments are manufactured with a small integrated shaft. Thus the gears

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can be precisely clamped to each of the general shafts by an integrated collet chuck

(cf. figure 2). A high precision collet chuck, which is also used in micro milling

machines for high accuracy machine tools, is inserted into a grinded conic hole at the

top side of the shafts and fixed by a clamping nut (true running accuracy < 2 µm).

Figure 1: Overview of the measurement setup

Each shaft is grinded and mounted by means of two high precision spindle ball

bearings which are fixed into grinded holes in the bearing bases as illustrated in

figure 2. Consisting of ceramic balls, the bearings have a very high stiffness and wear

resistance. Both bearings are preloaded in a duplex bearing (DB) arrangement to

enhance their running smoothness (true running accuracy < 2.5 µm) by means of a

locking ring, a distance sleeve, a clamping lid, a clamping sleeve and a groove nut.

In order to adjust the center distance of the micro gears to each other and to align

their rotational axes in parallel, the motor and the respective bearing base are

mounted on a 4-axes manipulator. It consists of two lateral units (positioning

accuracy < 3 µm) and an angular unit as well as a goniometer on top of those

(cf. figure 1). Further positioning units are not necessary for the adjustment of the

gears in lifetime experiments. The rig can be mounted to a coordinate measuring

machine so that the center distance and the axial orientation of the micro gears to

each other can be determined precisely (length measuring error < 2.4 µm) by

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measuring the grinded shaft cylinder between the bearing and the collet chuck

(cf. figure 2). Based on this, the manipulator system can be adjusted.

Figure 2: Bearings, shaft and gear fixing concept in the measurement setup

4 Summary and outlook

In this article, an experimental setup to conduct abrasive experiments of micro gears

under clearly defined conditions was presented. Speed, torque and the alignment of

the gears to each other can be controlled precisely. Besides, suitable concepts for the

fixing of the micro gears and the bearings of the shafts have been developed.

Currently, the rig is physically assembed and a mechanism to detect the time of the

gear defect is developed. Upcoming, the lifetime experiments will be started in order

to deduce the aforementioned lifetime model of micro gears dependant of their

shape deviations and material structure.

References:

[1] VDI guideline 2731: Microgears - basic principles. Berlin, Beuth, 2009

[2] M. Beier, Lebensdaueruntersuchungen an feinwerktechnischen Planetenrad-

getrieben mit Kunststoffverzahnung. Dissertation, University of Stuttgart, 2010

[3] B.-R. Hoehn, P. Oster C. Braykoff, Scuffing and Wear Load-Carrying Capacity

of Fine-Module Gears. International Conference on Gears, pp. 1295-307, 2010

[4] A. Albers, N. Burkardt, T. Deigendesch, C. Ellmer, S. Hauser, Validation of

micromechanical systems. Microsystem Technologies 14, pp. 1481-1485

[5] S. Hauser, Concepts for the validation of geometrical features of micro gearings

and gear boxes. Dissertation, University of Karlsruhe (TH), 2007

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3D-Reconstruction of Microstructures on Cylinder Liners

F. Engelke, M. Kästner, E. Reithmeier

Institute of Measurement and Automatic Control – Leibniz Universität Hannover

[email protected]

Abstract

Microstructuring of cylinder liners used in internal combustion engines is researched

to reduce friction and thereby save fuel and increase engine longevity. To fully

describe the effects of microstructuring on roughness and function of the surface, a

3D-acquisition, reconstruction and evaluation of the geometric features of the

microstructures is necessary. We developed methods using optical white light sensors

to reconstruct the 3D surface using subsequent measurements under varying surface

to sensor angles. We improved fourier based image alignment techniques and applied

them for the registration of partial images to form larger high resolution

measurements which are then aligned with measurements using different surface to

sensor angles. These data were used to form a 3D dataset of the microstructured

surface including measured undercuts.

1 Introduction

Currently the microstructuring of cylinder liners used in internal combustion engines

has come to the focus of a number of research groups. The effects of the

microstructures need investigation to characterize the caused reduction of friction

with the result of increased engine longevity and reduced fuel consumption [1,2]. To

fully describe these microstructures a 3D-reconstruction and subsequent analysis is

necessary, especially to find undercuts which cannot be detected using light

microscope measurements without variation of the measurement angle. Other

approaches for 3D reconstruction of microstructures describe methods for serial

sectioning, scanning electron microscopy and micro stereo vision respectively [3-5].

2 Sensors and Measurement

Methods were developed to measure microstructured surfaces under varying sensor to

surface angles using white light interferometry (WLI) for surface samples. Also a

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confocal chromatic sensor (CCS) which allows for measurements inside of cylinder

liners due to application of a specially designed sensor head is used to acquire

measurements under varying surface to sensor angles. For each angle either a number

of high resolution measurements by WLI are stitched together or a large number of

point measurements by CCS using a coordinate measuring machine (CMM) are fused

with positioning data obtained from the CMM by interpolation to form a heightmap

of the surface. In both cases the resulting measurements are repeated for different

surface to sensor angles to acquire the necessary data to form a real 3D data set from

2.5D data.

Figure 1: 1D-feature functions for overlap estimation (left), stitched microstructure

(top right), profile image of undercut specimen (bottom right)

3 Image alignment using 1D feature functions

The measurement by WLI was performed using a numerical aperture of 0.55 and a

measurement area of 0.25 by 0.19 mm². The length of the microstructures lies

between 1.2 and 1.5 mm along the cross section, which made multiple measurements

necessary to fuse them to form a single, high resolution measurement. New methods

for image alignment were developed to decrease the needed overlap between the

different measurements. Image alignment methods using correlation by fast fourier

transform are well known for robustness against noise and for being computationally

cheap compared to cross correlation [6], but require large overlap areas. We

elaborated on ideas presented in [7] to use 1D functions based on integral height

values along pixel columns. To achieve robustness against translational errors we also

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used other 1D functions based on the height values along the columns. We used the

standard deviation, the maximum, the minimum and the median. As these functions

are mathematically independent of one another we are able to gain additional

accuracy in predicting the overlap area by correlating these five functions. The

positional information is then used to restrict the areas of the input measurements for

correlation by fast fourier transform (FFT). By this method we are able to decrease

the necessary overlap for a successful alignment down to 7%.

4 Fusion of 2.5D data to full 3D data

To reconstruct the volumetric information of the microstructures the different

measurements are converted to point clouds, which are rotated by a least squares

algorithm, aligning the reference planes of the measurements. Because usual plane

fits do not compensate for elongation of the measurement grid due to rotation, we use

the transformed point clouds, which are interpolated to gain 2.5D data sets. The

structures are segmented using histogram based methods to temporarily remove the

structure. This is necessary as the data of the microstructures is dependent on the

measurement angle while the reference plane is not affected by rotation. The planes

with the micro structure edges are registered using the method described in section 3.

Each point cloud is used to generate a binary 3D matrix in which ‘0’-voxels represent

the object material. The 3D datasets are fused by application of a logical ‘or’.

5 Application for undercut detection

Undercuts are detected by three different methods. First, the number of true voxels

below the surface height, identified by histogram based methods, is counted for each

image and compared to the fused version which gives an indication on the existence

of undercuts as well as volume information, second, the true voxels below the

segmented surface are counted and third, the positions of specific undercuts are

detected by morphological 3D-thinning. This allows for a volumetric analysis of the

microstructures, which benefits the characterization for oil retention capacity.

6 Conclusion

We developed methods by which steep slopes and undercuts can be detected by use

of single optical sensors. To achieve this we developed new methods for image

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registration based on 1D-functions. We showed how undercuts can be evaluated in

volumetric binary data.

7 Acknowledgement

This work was partially financed by the “Deutsche Forschungsgemeinschaft” as part

of the project “Datenfusion optisch flächenhaft erfasster Mikrotopografien mit

Bezugsebene” (Data fusion of optically measured micro topographies with plane of

reference).

References:

[1] Tomanik, E. (2008). Friction and wear bench tests of different engine liner

surface finishes. Tribology International, 41(11), 1032-1038.

[2] Yin, B., Li, X., Fu, Y., & Yun, W. (2012). Effect of laser textured dimples

on the lubrication performance of cylinder liner in diesel engine.

Lubrication Science. 24(7), 293-312

[3] Lee, S. G., Gokhale, A. M., & Sreeranganathan, A. (2006). Reconstruction

and visualization of complex 3D pore morphologies in a high-pressure

die-cast magnesium alloy. Materials Science and Engineering: A, 427(1),

92-98.

[4] Samak, D., Fischer, A., & Rittel, D. (2007). 3D reconstruction and

visualization of microstructure surfaces from 2D images. CIRP Annals-

Manufacturing Technology, 56(1), 149-152.

[5] Wang, Y., & Liu, J. (2009, August). 3D shape reconstruction of

microstructures via micro stereovision. In Mechatronics and Automation,

2009. ICMA 2009. International Conference on (pp. 1861-1865). IEEE.

[6] Guizar-Sicairos, M., Thurman, S. T., & Fienup, J. R. (2008). Efficient

subpixel image registration algorithms. Optics letters, 33(2), 156-158.

[7] Guthier, B., Kopf, S., Wichtlhuber, M., & Effelsberg, W. (2012, April).

Parallel algorithms for histogram-based image registration. In Systems,

Signals and Image Processing (IWSSIP), 2012 19th International

Conference on (pp. 172-175). IEEE.

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A self-calibration method for the error mapping of a 2D

precision sensor

M. Valenzuela1, M. Torralba2, J.A. Albajez1, J.A. Yagüe1, J.J. Aguilar1 1I3A, University of Zaragoza, Spain 2Centro Universitario de la Defensa, Zaragoza, Spain

[email protected]

Abstract

In this paper two calibration techniques for the error mapping of a 2D sensor – a cross

grid encoder – are presented: a mathematical correction to assess the squareness

errors and a self-calibration of the cross grid encoder itself. The calibration setup

includes a metrology frame made of Zerodur®, a very low thermal expansion

coefficient material in order to reduce thermal errors. Additionally, an analysis by

means of a finite element analysis software has been carried out for an adequate

design of the set up. Finally, uncertainty values for the 2D cross-grid encoder system

are estimated.

1 Introduction

2D cross-grid encoders are very suitable to be used not only in the machine tool area

but also in metrology precision applications, such as, coordinate measuring machines

characterization and 2D optoelectronic sensors calibration. However, the calibration

of a 2D cross grid encoder is done by the manufacturer just in their main two axes

separately, which can be an accuracy limitation in some high precision applications.

To calibrate the whole area of the grid a calibration technique was proposed in a

previous work where the grid was used as a squareness reference [1]. But if this

cross-grid encoder is not perpendicular enough this error can also be a source of

influence in the final uncertainty of its calibration. To solve these calibration

problems, two different techniques are proposed in this work. The first one involves a

correction in the mathematical model presented in [1] to assess the squareness error

of the 2D cross-grid encoder. The second one is the application of a self-calibration

technique that includes the lack of squareness of the 2D cross-grid encoder. Besides,

in order to meet nanometer accuracy in this procedure, the use of very low thermal

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expansion materials together with a controlled environment (temperature, humidity

and pressure) are necessary.

2 Zerodur metrology frame design

The metrology frame used in this work is made of Zerodur® due to its very low

coefficient of thermal expansion in addition to its light weight which is necessary for

the assembly. This frame consists of two parts, a base and a top. Since Zerodur is a

brittle material, it has to be assembled without any direct contact with some metals or

any sharp material that can cause a micro fissure in the Zerodur. Different

preliminary tests involving a first design of a piece of aluminium (instead of

Zerodur), one or three screws with plastic or rubber washers to fix the parts and two

capacitive sensors were used to measure the stability of the system, as shown in

Figure 1 (a). After analysing the results it was decided that the best option to our

application consisted of using three screws and plastic washers. Once the couple of

Zerodur top plate was disposed, an optimum Zerodur plate form was designed and

analysed using Ansys Workbench software, taking into account that the top plate is

exposed to compression and tensile stresses as detailed in Figure 1 (b). The results

show that Zerodur top plate geometry is adequate to withstand tension and

compression stresses.

Figure 1: a) Stability test setup; b) Tension and compression stress acting on Zerodur

3 Cross-grid encoder calibration methods

The proposed setup shown in Figure 2 is mounted on a 2D moving table and it

includes the metrology frame described above, the 2D sensor to be calibrated (a

Heidenhain grid encoder KGM 181 with nanometer resolution, comprising a grid

plate with waffle type graduation and a scanning head) and a 2D laser encoder system

a) b)

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as a reference instrument (a Renishaw laser encoder RLE dual axis system with

nanometer resolution that includes RCU environmental compensation units).

Figure 2: a) proposed setup; b) angular misalignments of KGM.

As mentioned before, a mathematical model to calibrate the KGM was proposed in a

previous work [1], where a perfect perpendicularity in the cross-grid encoder was

assumed. Nevertheless, the grid encoder could have squareness errors that can

influence the final uncertainty of its calibration. One way to approach this problem

comes by assuming that all X lines and Y lines in the KGM grid are parallel but not

perpendicular between them, as shown in Figure 2b. If this squareness error is

included in the mathematical model presented in [1], then the new mathematical

model that relates the cross-grid encoder and the laser read-outs is as follows:

' ' '

' ' '

cos cos( ) cos sin( ) cos / cos

cos sin( ) cos cos( ) cos / cos

KGMX X X Y X X L X X

KGMX Y Y Y Y Y L Y Y

X X

Y Y

(1)

Another proposed way to fully calibrate the KGM would include the use of the self-

calibration method presented in [2]. The KGM error map is taken out from three

different views of the KGM, a normal view (view 0), another one rotated 180º with

respect to view 1 (view 1) and a translated view in the positive X axis direction (view

2). In each view the measurement deviation from X and Y KGM position and the

nominal position of laser system are denoted as:

0, , , , , , , 0, , ,x m n x m n x m n x m nV G L E ; 0, , , , , , , 0, , ,y m n y m n y m n x m nV G L E for view 0 (2)

1, , , , , , , 1, , ,x m n x m n x m n x m nV G L E ; 1, , , , , , , 1, , ,y m n y m n y m n x m nV G L E for view 1 (3)

2, , , , , , 1, 2, , ,x m n x m n x m n x m nV G L E ; 2, , , , , , 1, 2, , ,y m n y m n y m n x m nV G L E for view 2 (4)

where m = n = -(N-1)/2…(N-1)/2, Gx and Gy are the KGM error function in the

Cartesian space grid of 13 x 13 points (N x N) in an area of 60 mm x 60 mm (L x L)

b)

X laser head

Y laser head

Y laser mirror X laser mirror

KGM 181

a)

RCU units

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around the centre of the KGM. Lx and Ly are the measurement deviation of the grid of

the laser from the initial Cartesian space grid. Ex and Ey are the misalignment errors

between KGM and Laser axes.

4 Uncertainty analysis

Once the error maps of the grid encoder are calculated, an uncertainty analysis of this

2D sensor is carried out, as follows:

22 2 2 2.

, ,

.

( 2) Cal KGMLaser Error Residual T Resolution KGM

Cal KGM

SU k k u u u u

n (5)

where Laseru is the reference 2D laser uncertainty,

.Cal KGMS is the standard deviation of

the KGM calibration, .Cal KGMn is the number of data in calibration procedure,

,Error Residualu is the final error after alignment uncertainty, Tu is the uncertainty of

expansion/contraction of KGM due to small changes in temperature at test with

constant temperature and ,Resolution KGMu is the KGM resolution uncertainty. The values

obtained are between 300 and 400 nm both for X and Y axes.

5 Conclusion

In this work, two different calibration techniques for a 2D cross grid encoder are

presented using the same thermally stable setup and a 2D laser system as a reference

system. Finally, an uncertainty analysis of this 2D encoder is described.

Acknowledgements

This project was funded by Spanish government project DPI2010-21629 “NanoPla”.

Appreciation to DGEST which sponsored the first author.

References:

[1] J.A. Yagüe-Fabra, M. Valenzuela, J.A. Albajez, J.J. Aguilar, A thermally-stable

setup and calibration technique for 2D sensors, CIRP Annals, Manufacturing

Technology 60 (2011) 547-550.

[2] J. Ye, M. Takac, C.N. Berglund, G. Owen, R.F. Pease, An exact algorithm for

self-calibration of two-dimensional precision metrology stages, Precision

Engineering 20 (1997), 16-32.

139


Reaming in Microscale of Titanium and Titanium Alloys

D. Biermann, J. Schlenker

Department of Machining Technology, Technische Universität Dortmund, Germany

[email protected]

Abstract

The presented studies deal with adapting the process of microreaming pure titanium

and TiAl6V4 using tool diameters of one millimetre. The main focus is on generating

a high surface quality with low tool wear.

1 Introduction

Micromachining is considered to be a suitable technique for cost-efficient

manufacturing of microstructured parts in small or medium batch sizes such as molds

for micro molding processes. For the increasing market of microstructured

components micro holes with high quality are needed. In comparison to macroscale

drilling and reaming new problems appear, such as low tool stiffness. This can lead to

tool deflection or even to tool breakage. Because of this, high requirements

concerning surface quality and accuracy of shape can´t be achieved. An additional

process like reaming is expensive and complex, so, e.g. in [1, 2] the parameters are

examined for high quality of the hole and minor micro burr formation when drilling

brass, titanium, and aluminum.

Presently, there is an increasing trend to apply micro components and implants made

of high strength materials such as titanium and titanium alloys. The main application

area for micro components made of titanium alloys is the medical technology because

of their favourable physical and mechanical properties such as low density and high

corrosion resistance. However, titanium and titanium alloys belong to the group of

materials that are hard to machine due to their low thermal conductivity, low elastic

modulus, and high yield strength, causing a high thermal and mechanical load on the

cutting tool. In macroscale, there are already many publications dealing with

machining of these materials [3]. But in microscale basic researches are needed.

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2 Experiments

In the present analysis optimal parameter values are investigated in order to generate

a high surface quality, high accuracy of shape, low tool wear, and minor mechanical

stresses. The focus of this paper lies on the analyse of the influence of process

parameters (vc = 3 – 30 m/min and fz = 0.005 – 0.1 mm), radial allowance

(ar = 0.01 mm, 0.02 mm, and 0.03 mm), and lubrication system (dry, minimum

quantity, flood, and dipping lubrication) on the cutting process. The drill holes used

here are made by drill bits with diameter of d = 1 mm. The tested reamers have

diameters of d = 1.01 mm, d = 1.02 mm, and d = 1.03 mm to investigate the influence

of the width of the cut on process results. Furthermore, a comparison of the two most

common titanium alloys pure Titanium (Ti Grade 1) with 159 HV 0.02 and TiAl6V4

(Ti Grade 5) with 382 HV 0.02 is carried out. Also different lubrication systems (dry,

minimal quantity lubrication (MQL), flood lubrication, and dip lubrication) and their

influence are investigated.

3 Results and Discussion

3.1 Parameter

To analyse the influence of cutting speed vc and feed per tooth fz on the microreaming

process, these parameters were varied using design of experiments. The results prove

the increasing of hole diameters and their deviation with the cutting speed (cf. Figure

1). But the mechanical stresses are independent of the speed. Neither cutting speed vc

nor feed per tooth fz affect the surface quality of the bore wall. The important

influence of cutting speed on hole diameter is caused by the increasing tool deflection

in micromachining.

Figure 1: Influence of cutting speed on diameter and that variation

141


3.2 Radial allowance

Within the investigation of the influence of the radial allowance on the process, the

results of three different reamer diameters d1 = 1.01 mm, d2 = 1.02 mm, and d3 =

1.03 mm were compared. The use of the reamer d1 = 1.01 mm could not improve the

surface quality in comparison to drilled quality, and the tool with d3 = 1.03 mm

failed after only a few millimeter machining. The evaluation of the mechanical stress

shows that the force in the feed direction increases with increasing radial allowance

(cf. Figure 2). It can be concluded that when using the smallest reamer diameter the

minimum chip thickness was not achieved. So it leads to almost no material

removal. In contrast, the largest diameter achieves the best surface quality, but for an

economic tool life the mechanical stresses are too high. In summary the optimal

diameter is d2 = 1.02 mm.

Figure 2: Influence of reamer diameter on mechanical stresses

3.3 Lubrication System

Evaluating the mechanical stresses and the resulting diameters, the values of dry

machining and minimum quantity lubrication, and the dip and flood lubrication

appear to be very close together. For the first two, the process creates a larger hole

diameter and higher mechanical stresses in the feed direction (cf. Figure 3). When

reaming, the slight lubrication film of the MQL does not sufficiently wet the cutting

edges and the chip flutes. To gain the positive influence of the lubricant, the use of

dip or flood lubrication is necessary.

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Figure 3: Influence of lubrication systems on mechanical stresses

4 Conclusions

The results show that a lower cutting speed is better for reaming pure titanium and

Ti6Al4V. The use of reamers with diameter of d2 = 1.02 mm is to recommend for

machining drills with diameter of d1 = 1 mm. Dry and minimum quantity lubrication

is not as effective as dip and flood lubrication.

Acknowledgements

The presented investigations were supported by RWTÜV foundation within the

project “Micromachining of titanium and titanium alloys”.

References

[1] Denkena, B.; Hoffmeister W., H.; Reichstein, M.; Illenseer, S.:

Mikrozerspanung. In: Hesselbach, J.; Wrege, J. (Hrsg.): Kolloquium

Mikroproduktion, Braunschweig, 2003, S. 65-74

[2] Biermann, D.; Schlenker, J.: Studies of Microdrilling Titanium and Titanium

Alloys. 12th International Conference of the European Society for Precision

Engineering and Nanotechnology, Volume II, 3.6.-8.6. 2012, Stockholm,

Schweden, Spaan, H.; Shore, P.; Burke, T. (Hrsg.), S. 233-236

[3] Ezugwu, E.O.; Wang, Z.M.: Titanium alloys and their machinability – a

review, Journal of Materials Processing Technology 68 (1997) 262-274, 1995

143


Investigation of stylus tip-size effects in surface contact

profilometry

Khalid T. Althagafy¹’², D G Chetwynd¹

¹School of Engineering, University of Warwick, Coventry CV4 7AL, UK

²Umm AlQura University, Makkah, 21955, Saudi Arabia

[email protected]

Abstract

This paper presents refinements to 3D simulations of stylus effects in

microtopography measurements. It briefly reviews how statistically richer data can be

obtained by extending basic kinematic models, perhaps providing steps towards more

sophisticated modelling of the contact process. After a few notes about the new

simulation scheme, some illustrative results concentrate on idealized styli operating

beyond the limit of their expected resolving power.

1 Introduction: stylus simulation

It is almost impossible to determine directly the complex interactions that occur

between a profilometer stylus and a surface during a measurement of surface micro-

topography. There is, however, a long history (dating back to the obsolete E-system

of reference lines) of simulating the loci of circles, and latterly spheres, rolling on

rough topography both as an evaluation tool and in attempts at ‘stylus deconvolution’.

Comparison between real results and such simulations can reveal indirect evidence

about the behaviour in the contact region, although current models are hardly

sophisticated enough to make much impact. They are purely kinematic models,

assuming perfect guides and rigid objects and are actually a class of morphological

filter (specifically, the centre locus is dilation operation) [1]. Published work

concentrates heavily on the effect of the ‘measured surface’, extended in one case to

report some statistics on the contact to the stylus [2]. The current research takes this

further, motivated by its potential to cast light onto the in-plane uncertainty of stylus

measurements, suggest ways to improve comparisons between instruments or lead

eventually to methods for self-diagnosis of stylus wear.

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Figure 1: Conical, spherical and pyramid (5µm) Stylus tip shapes

A new simulation gathers full data about stylus contact related to its location upon the

surface. It introduces a threshold process [3] by which the kinematic condition is

violated in small increments and the growth of resulting ‘contact areas’ recorded.

This is intended to give insight, into sensitivity to instrument noise, repeatability, etc.

It also has potential for modelling the contact process, by approximating stiff but non-

rigid contact using relaxation techniques. The contact modelling is implemented in

MATLAB® which is interfaced with the commercial topographic analysis software

SPIP in order to provide a standard for parameter evaluation and comparison, and to

translate between different instrument data formats and MATLAB arrays [4]. The

surface could be any set of data representing a real or an arbitrary surface. Also, the

stylus could be any set of data representing a real or an arbitrary stylus shape. Both

data sets are dealt with as arrays.

2 Styli and fine surface structure

This report concentrates on study of the sensitivity to stylus shape and condition

when detecting features of real surfaces at the very limits of conventional

profilometer capabilities. It therefore uses relatively small arrays with a grid sampling

interval of 0.1 µm. Three ideal computer generated stylus tips with different shapes

have been used: conical, pyramid and spherical (Figure 1). The tip radius and heights

(for the conical and pyramid shape) are 3 µm and 5 µm. The tip angles of the conical

and pyramid shapes are 90⁰. Each tip has been used in its perfect shape and with a

quite severe truncation at 2µm below the original tip. The spherical tip offers a full

hemisphere, not the more usual blend into a cone, allowing estimation of flank

contact for different cone angles. Many trials were run of the basic stylus shapes over

simple computer generated surfaces, such as single or clustered delta functions. These

can rapidly identify major bugs and increase confidence that no subtle errors remain

in the simulation routines.

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3 Results

Surface maps of the fine structure of ground steel surfaces were measured by Atomic

Force Microscopy (AFM) to ensure high lateral resolution. The data collected by the

AFM were checked for missing data and interpolated by the SPIP software. Figure 2

shows a typical example.

Table 1: Error of the roughness parameters of different tips on the ground surface

Roughness

Parameter

Sa Sq Sy Ssk Sku

458 nm 481.5 nm 920 nm 1.135 1.37

Stylus Tip Size Shape %Error=100x(Measured Value - Actual Value )/ Actual Value

Pyramid 3µm Perfect 0.00 0.00 0.00 0.00 0.00

2µ Worn -2.1 -1.5 -7 -1.8 -3.7

5µm Perfect 0.00 0.00 -0.02 0.00 0.00

2µ Worn -3.62 -3.44 -6.88 -0.877 -1.43

Sphere 3µm Perfect -1.09 -1.4 -17.2 -0.88 -0.72

2µ Worn -34 -29 -11 -7.96 -16.7

5µm Perfect -6.76 -6.76 -1.6 -18.4 -0.877

2µ Worn -65.34 -56.34 -29.9 -10.57 -20

Cone 3µm Perfect 0.00 0.00 0.00 0.00 0.00

2µ Worn -2 -1.4 -7.39 -1.7 -3.64

5µm Perfect 0.00 0.00 0.00 0.00 0.00

2µ Worn -3.38 -3.34 -7.45 -0.877 -1.4

Surface maps were scanned in simulation by the set of 3 µm and 5 µm styli (which

would normally be considered too large for the task). Figure 3a and 3b shows two

illustrative profiles taken from scans on the data in figure 2. The 3 µm hemisphere

does remarkably well on local detail. Table 1 shows the percentage error of selected

Figure 2: 20x20µm ground

surface on 0.2µm sample grid taken by AFM

(b) (a)

Figure 3: Pprofiles taken from scans across the data in

figure2 using (a) 90⁰ Pyramid stylus (b) 5 µm

Hemisphere stylus.

146


roughness parameters of different outputs when scanning the same ground shape

with the different tips.

As expected, the maximum deviation occurs when using the 5 µm spherical tip with

2 µm truncation, but many cases show only small errors. The simulations show that

in most cases contact do not occur at the central point of the stylus, even with

idealized shapes other than perfectly shape ones. With the spherical tip, the mean

position of the contact is close to the centerline, while its slandered deviation is

about 0.5 µm on 3 µm radiuses. It tends to the radius of the flat on truncated ones.

Initial evidence shows that examination of the contact patterns as threshold increases

can identify the intensity with which different asperity regions interact with the

stylus. For example, a 5 nm threshold caused little change in contact sizes from the

kinematic point, but 50 nm caused them to grow asymmetrically, eventually picking

out the major structures of the surface.

4 Conclusion

A new simulation program has been developed and used to examine the measuring

fine structure of real surfaces by the stylus method. Although able to scan any

arbitrary surface with any arbitrary stylus shape, the results given here use idealized

styli and ‘real’ ground steel surfaces.

The simulations have naturally confirmed that the stylus geometry and size can have

a significant effect on most roughness parameters of the measured surface in 3D.

The surprising feature of them, worthy of greater investigation, is how insensitive to

major changes in stylus condition, some of the popular parameters are, even when

dealing with very fine structure within localized areas of a ground surface.

References

[1]Muralikrishnan B, Raja J. Computational Surface and Roundness Metrology.

Springer-Verlag London. 2009; Ch 2, 8

[2]Dowidar HAM, Chetwynd DG. Distribution of surface contacts on a simulated

probe tip. FLM Delbressine et al. (eds) Proc. 3rd International euspen Conference,

Eindhoven. May 2002; 757-760.

[3]Khalid T. Althagafy, Chetwynd DG. Simulation of Stylus Contact Patterns in

Profilometry.Styli. Proc. 26th ASPE Annual Meeting, Denver, US, November 2011.

[4]Khalid T. Althagafy, Chetwynd DG. Simulation Studies of Sub-micrometer

Contact of Topography Styli. Proc.27th ASPE Annual Meeting,San Diego,US, 2012.

147


ISO Compliant Reference Artefacts for the Verification of

Focus Varation-based Optical Micro-co-ordinate Measuring

Machines

F. Hiersemenzel1, J. D. Claverley2; J. Singh1, J. N. Petzing1, F. Helmli3, R. K. Leach2

1Loughborough University, Loughborough, UK;2National Physical Laboratory, Teddington, UK;3Alicona Imaging GmbH, Graz, Austria

[email protected]

Abstract

Demand for micro-co-ordinate measuring machines (micro-CMMs) within industry is

increasing due to the need for accurate measurement of the geometry of small-scale

objects. Optical micro-CMMs have the advantage over traditional stylus-based

CMMs of being non-contact instruments, and have the potential to acquire large

amounts of data, with high resolution, in a relatively short period of time. The focus

variation (FV) technique is typically used for surface topography measurement, but

has the potential to be implemented as a sensor technology for optical micro-CMMs.

Exploring the possibility of the FV technique as part of an optical micro-CMM

requires a robust performance verification of the instrument and measuring

procedure, using material measures that are traceable to the definition of the metre.

This paper proposes a design for a calibration artefact that is suited to volumetric

verification for micro-CMMs based on the FV technique and recognizes recent

developments of ISO 10360.

1 Introduction

The ISO 10360 specification standard for acceptance testing and verification of

CMMs has several parts, all specific to different groups of instruments and

configurations. Each section of ISO 10360 identifies methods and artefacts best

suited for the acceptance testing and verification of each group and configuration.

ISO/DIS 10360-8.2 [1] (due for ISO/FDIS publication in 2013), is a verification

standard written for CMMs with optical distance sensors. There are four main parts to

the acceptance and re-verification tests: length measurement error, probing form error

measurement, probing size error measurement and flat form error measurement. The

148


probing form and size error tests require a calibrated reference sphere that has a

diameter of at least 10 mm. Existing FV surface topography measuring instruments,

when used as micro-CMMs, are potentially covered by this standard but the

recommended minimum size of the calibration sphere is too large to fully fit within

one field of view of a typical system. Most surface topography measuring instruments

similar in operation to FV systems, such as confocal microscopes and coherent

scanning interferometers are, therefore, also currently excluded from the application

of this standard by default, unless the user, and the instrument manufacturer, can

agree to use a smaller calibrated reference sphere for the assessment of the probing

size and form error. A prototype FV-based optical micro-CMM should, therefore, be

verified with calibrated reference spheres of similar size to objects for which the

technique has been designed to measure.

Numerous calibration artefacts exist for micro-CMMs (a review is being written by

NPL); however, most do not fulfil all the criteria required for a FV instrument. For

instance, artefact surfaces are often too smooth to be measured by the FV technique,

and the dimensional layout of the artefact is not suitable for an optical instrument

with a short stand-off distance. Calibration artefacts have to be designed taking into

consideration the conditions for which the instrument performs best, specific

requirements for the technology used, whilst also maintaining a traceability chain to

the definition of the metre.

2 Novel verification artefact

A novel verification artefact developed for the prototype FV-based optical micro-

CMM is composed of multiple small-scale spheres mounted in tiered equally-spaced

conical holes. The artefact is specifically designed to evaluate: probing size error,

probe form error, and system dimensional accuracy, compliant to ISO/DIS 10360-8.2.

A photograph of the artefact is shown in Figure 1.

Reference spheres are suitable components for this verification artefact because they

do not have sharp edges (which may cause object illumination problems) and are

geometrically ideal in terms of data fitting and modelling. The considerations for FV

include rough surface specifications (minimum Ra ≈ 30 nm), surface slope, and

accessibility for high magnification lenses (which generally have smaller standoff

distances, in the order of several millimetres).

149


The spheres are aligned on body diagonals and face diagonals that always include the

z direction. The sphere diameter chosen for the artefact is 1 mm, although 0.5 mm

and 2 mm diameter spheres have also been considered [2]. Four materials have been

tested for their suitability for the reference spheres and of these, stainless steel has

been chosen for the prototype artefact, but silicon nitride has also been found suitable.

[2]

Figure 1: Verification artefact for an optical micro-CMM based on FV.

3 Distance measurement between two spheres

The distance between two consecutive spheres (on the same plane) was measured

with a FV instrument using a 50× magnification objective lens, two separate

(unstitched) fields of view and the same co-ordinate system. In order to be able to

measure as much of the spheres as possible, a ring light and polarizer were used.

These serve to increase the illumination aperture of the system, and help improve the

detection of scattered light from high angle surfaces. Low lateral and vertical

resolutions, 2.93 μm and 0.68 μm respectively, were chosen to minimise the

measurement duration. Applying a robust sphere fitting algorithm [3] to the

measurement results gives the 3D co-ordinates for the centres of the spheres. From

these, the distance between the sphere centres can be calculated. This measurement

procedure was repeated three times.

In order to have a comparison to the performance of the FV instrument, the same two

spheres have been measured on a high accuracy traditional contact CMM (MPE =

(0.7+L/600) µm where L is the nominal distance measured in millimetres), measuring

and fitting to each sphere using five data points.

150


4 Experimental results and conclusions

An example of the result for the separation between two sphere centres measured

three times using the FV instrument was 7.122 mm (standard error 0.001 mm), whilst

the repeated measurement result for the same spheres using the CMM was 7.112 mm

(standard error 0.000 06 mm). The measurement of the FV instrument tends to have

higher linear distance values and larger standard deviation values, potentially because

the instrument is primarily designed to rely on post process image-stitching. Further

sphere combinations have been, and are being, evaluated.

Experimental work is showing the potential of a FV-based instrument to function as

an optical micro-CMM. The initial results suggest that the elements of the procedures

detailed in ISO/DIS 10360-8.2 can be applied to optical micro-CMMs, thereby

providing a traceable verification route to the metre.

The calibration artefact, as shown in Figure 1, will undergo minor dimensional

changes in order to optimise the calibration of the artefact using an established

contact micro-CMM. Further work will investigate the effect of lateral and vertical

resolution for dimensional measurements in the context of the novel calibration

artefact presented here, with cross comparison to traditional CMM data also being

completed. Consideration will also be given to issues and the merit of short term

health checking procedures of a FV-based optical micro-CMM, versus full

reverification. Health checking options with artefacts such as this will provide fast

estimation and monitoring of optical micro-CMM health. This recognizes similar

strategies identified in other parts of ISO 10360.

References

[1] ISO/DIS 10360-8.2, Geometrical Product Specifications (GPS) - Acceptance and

reverification test for coordinate measuring machines (CMM), Part 8.2: CMMs with

optical distance sensors, 2012, International Organization for Standardization

[2] F. Hiersemenzel et al. Development of a traceable verification route for optical

micro—CMMs, 10th International Conference on Laser Metrology, Machine Tool,

CMM & Robotic Performance - Lamdamap 2013, Kavli Royal Society International

Centre, Buckinghamshire, UK, March 2013

[3] A. B. Forbes, Robust circle and sphere fitting by least squares, Technical Report

DITC 153/89, National Physical Laboratory, Teddington, UK, 1989

151


Acoustic Emission-based micro milling tool contact

detection as an integrated machine tool function

E. Uhlmann, N. Raue, C. Gabriel

Chair for Manufacturing Technology, Technische Universität Berlin, Germany

[email protected]

Abstract

Various sensor-based methods for referencing tool and work piece position are

available for the contact detection between tool and work piece. In this article the

robustness of the acoustic emission-based contact detection will be investigated

which is a vital requirement for such a system in order to be integrated into machine

tools.

1 Introduction

One of the key challenges in micro milling is the precise and reliable detection of tool

position and work coordinate system. The system proposed in this paper uses a direct

method for tool contact detection. Via the application of an Acoustic Emission sensor

to the work piece holder of a micro milling machine tool, the contact detection can be

performed through the detection of acoustic emission generated by the contact of

work piece and rotating tool. A major advantage of such direct method is the

possibility of measuring the contact point with a rotating spindle, thus eliminating the

thermal error due to the axial expansion of the rotating spindle.

2 State of the art

Acoustic Emission (AE) is structure-borne sound of high frequency that is generated

by solids under mechanical strain. Sources of AE are plastic deformation, friction,

crack formation and material breakage [1]. In the context of the proposed application,

the AE signal is utilized to determine the moment of physical contact between a work

piece and the rotating milling tool in a vertical approach. This approach was already

proposed and experimentally evaluated by Bourne et al. [2] and Min et al. [3]. The

focus of this contribution, besides the further evaluation of the feasibility, is on the

integration of the AE-based contact detection into the control of the machine tool and

the accuracy of repetitive approach processes.

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3 Experimental setup

The proposed system was implemented on a prototypical 3-axes micro milling

machine with a Sycotec 4064 DC high speed spindle and a CNC of the type Beckhoff

TwinCAT. The Acoustic Emission sensor 8152B of the manufacturer Kistler is

inserted into a specially designed work piece fixture and thus mounted to the bottom

side of the work piece. After preamplification through a Kistler 5125B amplifier

module, the signal was acquired through a National Instruments USB-6351 data

acquisition device. TiAlNi-coated two flute end mills with a diameter of 500 µm were

applied in vertical approach experiments. Brass and steel (X38CrMoV5-1) were

examined as work piece materials. Before and after nine consecutive approaches, the

mills were analysed in a scanning electron microscope. The work piece surface was

optically scanned using the focus-variation-based surface metrology system Alicona

InfiniteFocus.

An incremental approach algorithm [4] was applied by implementing a semaphore

communication between the signal processing module and machine control. This

implies that the machine control initiates a downward movement of 1 µm after

receiving an explicit approval from the signal processing module. Subsequently, an

AE signal is acquired with a length of 500 samples at a sampling rate of 1 MHz and

the signal is filtered. If the signal energy of the processed signal is below a defined

threshold, no contact is detected and thus, an approval for the next downward motion

will be sent to the machine tool control.

Measurement

System

Measurement System Control System

Initiating Movement()

Confirm Movement

Signal Acquisition

Signal Test

Set Position()

Accept Position

No

Conta

ct

Work Piece

Sensor

DAQ Device

Preamplifier

Spindle

PLC

End Mill

Ax

is

Signal

Processing

Software

Control System

Software Interface

Figure 1: Experimental setup and sequence of incremental approach algorithm

153


In case of a threshold excess, the current z-coordinate is sent to the machine control

as the new reference point and an upward movement is initiated to prevent

unnecessary friction between tool and work piece.

4 Results

4.1 General results

The tools were affected by the contact event as it can be seen in Figure 2. Compared

to steel, brass as a work piece material imposed greater strains to the TiAlN-coated

tool.

New After nine

approaches on steel

After nine

approaches on brass

50 µm 50 µm 50 µm

Figure 2: Milling tool cutting edges before and after approach experiments

4.2 Reproducibility experiments

To evaluate the contact detection for the determination of the work piece surface the

approach was repeated 400 times in an automated procedure for each of the work

piece material (Figure 3). In both cases, each contact detection was performed

successfully. For compensating the inclination of the work piece surface due to

mounting error, a regression plane of the measured z-coordinates was determined.

......................

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......................

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......................

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......................

......................

......................

Experiments

Deviation

from Regression

Plane [µm]

Y Coordinate [mm]X Coordinate [mm]

-2-101234

Processing of machine control coordinates

02

46

810

1214

1618

20

02

46

810

1214

1618

20

Contact depth

Evaluation

......................

......................

......................

......................

......................

......................

......................

......................

............................................

......................

......................

......................

......................

......................

......................

......................

......................

......................

......................

Figure 3: Process sequence of the reproducibility experiments

154


For evaluating the significance of the tool contact detection the deviations of the z-

coordinates were compared to the cutting depths of ten points. Executing the

experiments with the brass alloy, the optical measurements indicate an average

cutting depth of 1.061 µm. The average deviation of the z-coordinate to the actual

work piece surface is 0.554 µm. These values are material specific and have to be

taken into account for determining the actual work piece surface with AE-based tool

contact detection.

n=400 Average Deviation of

the z coordinates from

Regression Plane [µm]

Range of the Deviation

of z coordinates from

Regression Plane [µm]

Average Deviation

from Regression Plane

to actual Work piece

Surface [µm] (n=10)

Steel 0.829 µm [-2.0, 3.4] 0.554

Brass 0.581 µm [-2.6, 2.5] 1.061

Table 1: Results of the Reproducibility Experiments

4.3 Discussion and further research

The presented work demonstrated the potential of an automatable AE-based contact

detection for micro milling machine tools. Further research will consider the contact

detection in x, y direction to determine the work piece position. Furthermore the

material-dependent depth variation of the surface damage should be investigated to

obtain a fixed offset between the regression plane and the actual work piece surface.

References:

[1] V. Zinkann, Der Spanbildungsvorgang als Acoustic-Emission-Quelle.

Aachen: Shaker, 1999.

[2] K. A. Bourne, M. B. G. Jun, S. G. Kapoor, and R. E. DeVor, “An Acoustic

Emission-Based Method for Determining Contact Between a Tool and Workpiece at

the Microscale,” J. Manuf. Sci. Eng, vol. 130, no. 3, p. 31101, 2008.

[3] S. Min, H. Sangermann, C. Mertens, and D. Dornfeld, “A study on initial

contact detection for precision micro-mold and surface generation of vertical side

walls in micromachining,” CIRP Annals - Manufacturing Technology, vol. 57, no.

1, pp. 109–112, 2008.

[4] S. Min, J. Lidde, N. Raue, and D. Dornfeld, “Acoustic emission based tool

contact detection for ultra-precision machining,” CIRP Annals - Manufacturing

Technology, vol. 60, no. 1, pp. 141–144, 2011.

155


Dimensional verification of high aspect ratio micro

structures using FIB-SEM

Y. Zhang1, H.N. Hansen1 1 Department of Mechanical Engineering, Technical University of Denmark

[email protected]

Abstract

Micro-structured surfaces are increasingly used for advanced functionality. In

particular, micro-structured polymer parts are interesting due to the manufacturing via

injection moulding. A micro-structured nickel surface was characterized by focussed

ion beam-scanning electron microscope (FIB-SEM) and then analysed by Spip®. The

micro features are circular holes 10µm in diameter and 20 µm deep, with a 20 µm

pitch. Various inspection methods were attempted to obtain dimensional information.

Due to the dimension, neither optical instrument nor atomic force microscope (AFM)

was capable to perform the measurement. Via FIB-SEM, the process was recorded

by images when slicing the sample layer by layer by ion-beam. In this way, the

dimension and the geometry of the holes are characterized.

1 Introduction

Micro polymer pillars arrays modify the wetting properties of the surface, for instance

previous research suggests that micro-structured surface can favour cells growth

when the pillars are patterned in certain ways [1], therefore it has a wide application

in bio-medical fields. Biocompatible polymers are used for this type of application.

The micro pillars array is a surface geometry; the dimension of the feature is typically

orders of magnitude smaller than the structured surface area [2]. A master geometry

is required for the replication of the micro pillars array. Lithographical methods are

often used to produce the master, i.e. the pattern of the pillars is introduced by

lithography and metal deposition (such as physical vapour deposition) with a mask.

Subsequently electroplating is used to create an insert for the moulding process.

In order to analyse the replication degree, it is necessary to characterize the mould

structure accurately. The nominate dimensions of the circular holes studied in this

paper are 10 µm in diameter and 20 µm deep, with a 20 µm pitch. For most types of

AFM it is beyond the measurement ability, unless a customized cantilever is used.

156



An optical microscope with Focus-Variation (Alicona®) was applied to measure the

depth of the holes, however, the bottom of the holes cannot be “observed” by the

microscope simply because the reflected light from the bottom was insufficient.

Figure 1 is the top view of the investigated surface, obtained by scanning electron

microscope (SEM). Similar to the result of an optical instrument, conventional SEM

has the difficulty to get sufficient information from inside the holes. When the sample

was tilted up to 30 degree, the surface of the inner wall was shown (Figure 2). But the

depth of the hole was still not illustrated.

Figure 1 A SEM image of the top of the

surface with micro holes.

Figure 2 The sample was tilted in SEM.

2 Conventional cross section measurement

Figure 3 epoxy-moulded of cross sections of the mould.

The result is influenced significantly by the alignment and cutting process. The same

scale is applied in these two images. Sample (b) was polished further based on sample

(a).

Another often used method to investigate the geometry is to make cross section of the

holes. The sample needs cutting from the side, then epoxy moulding in order to be

ground and polished. Due to the micro dimension of the structure, the obtained cross

(a) (b)

157


section is influenced significantly by the sample preparation process, such as the

alignment, the cutting step and the grinding step. Pictures (a) and (b) in Figure 3

show two different cross sections from the same sample; (b) was obtained by

polishing the sample in (a) 1 mm further down. Image (a) shows that the diameter of

the hole is approximately 6.5 µm, while image (b) shows the diameter of the hole is

8.5 µm. Neither of them confirm to the nominate value 10 µm. theoretically it is

possible to make such a cross section sample for every few micrometres of the

sample, presuming it is allowed by the cutting technique. However, it is not only time

consuming but also completely destructive, as a result this is not the first choice when

the sample material is expensive and the time schedule is tight.

2 Quanta 200 3D SEM FIB

Figure 4 FIB SEM milling process. From (a) to (h) the distance between two image is

1 µm, from (h) to (i) the distance is 2 µm.

(a) (b) (c)

(d) (e) (f)

(g) (h) (i)

158


The Quanta 200 3D was used in this study. It is a dual-beam scanning electron

microscope which combines normal SEM mode functionality. It uses a focussed ion

beam (FIB) for removing material by milling. In this study, the accelerating voltage

was 30 KV, using imaging detector Everhart-Thornley in high vacuum. The current

was 7 nA in the milling process.

A random hole was chosen to be observed. The side of the sample was positioned to

be vertical to the ion beam. A block of material was removed by ion beam until the

investigated hole was exposed. The sample was sliced from the side instead of from

the top, to avoid debris falling into holes. The hole was sliced with a step of 200 nm,

i.e. 200 nm thick materials were removed in each layer during the milling.

As the images in Figure 4 illustrate the hole was milled from the front to the back. (a)

and (b) show that the side wall of the hole was not perfectly perpendicular to the

milling beam direction; it was approximately 2.8 degree tilted. From image (c) the

contour of the hole is visible, as well as the structure on the inner wall.

The dimension was analysed by SPIP® using x-y scaling tool. Picture (e) in Figure 4

was used for this analysis, since it illustrates the central position of a hole. The

diameter is 9.7 ± 0.06 µm, the depth is 24.8 ± 0.06 µm considering the tilted angle.

3 Conclusion

A structured surface 10 µm in diameter and approximately 20 µm deep was measured

by conventional SEM and a FIB SEM. Due to the relatively high aspect ratio, only

FIB SEM can measure the depth of the hole by milling the hole from side. Compared

to conventional epoxy-moulded cross section method, FIB-SEM is relatively faster

and less destructive; meanwhile it requires much less preparation work.

References:

[1] E. Stankevicius et al, “Holographic lithography for biomedical applications “,

Proc. of SPIE, 2012; 843312

[2] H.N.Hansen et al. ”Replication of micro and nano surface geometries”, CIRP

ANN-MANUF TECHN, 2011; 60, 695-714

[3] Russell, P., D. Batchelor, and J. Thornton. "SEM and AFM: Complementary

Techniques for High Resolution Surface Investigations."

Acknowledgements

The authors would like to thank DTU Center for Electron Nanoscopy (CEN) for the

facilities support of The Quanta 200 3D dual-beam scanning electron microscope.

24,8µm

9,7µm

159


Setting-up kriging-based adaptive sampling in metrology

D. Romano1, R. Ascione2 1University of Cagliari, Italy 2ENEA, Italy

[email protected]

Abstract

Statistical sampling is a fundamental tool in science, and metrology is no exception.

The merit of a sample is its efficiency, i.e. a good trade-off between the information

collected and the sample size. Although the sample sites are ordinarily decided prior

to the measurements, a different option would be to select them one at a time. This

strategy is potentially more informative as the next site can be decided based also on

the measurements taken up to that time. The core of the method is to drive the next-

site selection by a non-parametric model known as kriging, namely a stationary

Gaussian stochastic process with a given autocorrelation structure [1,2]. The main

feature of this model is the ability to promptly reconfigure itself, changing the

pattern of the predictions and their uncertainty each time a new measurement comes

in. Since the model is re-estimated after each added point the sampling procedure is

an adaptive one. The next sampling site can be selected via a number of model-

based criteria, inspired by the principles of reducing prediction uncertainty or

optimizing an objective function, or a combination of the two.

The methodology has been applied by the authors [3,4] to design inspection plans

for measuring geometric errors using touch-probe Coordinate Measuring Machines

(CMM). Results showed that both the non adaptive statistical plans (Random, Latin

Hypercube sampling, uniform sampling) and two adaptive deterministic plans from

the literature were largely outperformed by the proposed plans both in terms of

accuracy and cost.

Here we further investigate on a number of important questions related to adaptive

kriging: which is the best trade-off between the number of adaptive and non-adaptive

points (the latter chosen according to uniform coverage), which next-site selection

criteria are more suitable to capturing extreme values of the signal in order to provide

a good estimate of the geometric errors.

160



1 Kriging modeling

Response y(x) is considered a realization of the Gaussian Process (GP) defined as

)()( xZxY , (1)

where is a constant and )(xZ is a GP with zero mean and stationary covariance:

;hRσhxYxYCovμxY Z

2,,)( . (2)

In (2)2

Zσ is the process variance, R the correlation function depending only on the

displacement vector h between any pair of points in the domain and on a parameters

set θ. The model defined by (1) and (2) is known as simple kriging. A flexible

choice for the GP correlation structure is the power exponential function:

2,1,20,01

1

exp

dipiθp

ihiθ

d

i

hR i , ; (3)

where θ=( θ1,…, θd, p1,…,pd), is a vector of unknown scale parameters (θ1,…, θd) and

smoothing parameters (p1,…,pd) respectively. Parameter θi describes how rapidly

correlation decays in direction i with increasing distance |hi|. Parameter p

i describes

the shape of the correlation decay (see Figure 1).

2 Criteria for next-point selection

We use three kinds of criteria: objective-specific, informative, and a combination of

the two. The objective-specific criterion (MaxF) point to maximize an objective

function, e.g. the signal itself or the geometric error. The two informative criteria

select next-point to inspect where the uncertainty of predictions by the current

kriging model is maximum (MaxPVar), and where prediction uncertainty weighted

by the distance from the nearest point already inspected is maximum, thus

promoting uniform coverage (MaxWPVar). Finally, composite criteria (switch rules)

are defined which select the next point as the one producing the maximum increase

of the objective function wrt the previous step; if no increase is possible the rule

switches to one of the two informative criteria (switch rule 1 (SR1): MaxF or

MaxPVar; switch rule 2 (SR2): MaxF or MaxWPVar). Prediction uncertainty is

evaluated empirically by using the Jackknife variance operator as it proved to

convey much more information than the so-called kriging variance. The latter, in

161


fact, holds only when the parameters of the correlation function are known, which is

hardly the case in practice.

= 0.1, p = 0.5 = 0.1, p = 1 = 0.1, p = 1.9

= 0.6, p = 0.5 = 0.6, p = 1 = 0.6, p = 1.9

= 0.7, p = 0.5 = 0.7, p = 1 = 0.7, p = 1.9

Figure 1: Nine one-dimensional signals (left) synthesized by as many GP models via

their power exponential correlation function (right)

3 Scope of the analysis

We look for the most effective settings of the adaptive procedure in terms of the size

of the initial non-adaptive sample, chosen as a Latin Hypercube one, and the

criterion for next-point selection, in view of maximizing the objective function, i.e.

the measured error after 40 inspected points (err_40). For this purpose we set up a

planned experiment whose factors (levels) are: range and shape parameters (0.1,

0.6, 0.7) and p (0.5, 1.0, 1.9), the size of the initial LHS sample, n (4 to 38, step 2),

and the criteria for next-point selection (five criteria, see section 2). Nine one-

dimensional signals (Figure 1), spanning a sizable interval of information content,

are generated by a random walk from each GP model obtained by crossing the three

levels of parameters and p. Ten replicates of the adaptive procedure are run.

4 Results

Main effects and two-factor interactions for the response err_40 are shown in Figure

2. The measured error after 40 inspected points is generally accurate, ranging from

90% to 100% of the true error. The shape parameter, p, of the correlation function

162


has a significant effect on the procedure's ability to getting close to the true error

while the range parameter is much less active. Low values of p, corresponding to

very noisy signals, make the error's estimate less accurate. The most interesting

result is that a 50%-50% allocation of the initial LHS points and the successive

adaptive points proves to be a good and robust (small variation of the response for n

in the interval from 8 to 34) choice for all the signals. The mean performance of the

five adaptive criteria seems similar and always superior to Random sampling and

LHS.

0,7

0,6

0,1

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98

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C3: M

axW

PVar

C2: M

axPV

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rr_4

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2

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F

100

90

80100

90

80100

90

80

0,1

0,6

0,7

teta

0,5

1,0

1,9

p

22

24

26

28

30

32

34

36

38

4

6

8

10

12

14

16

18

20

n

n

criterion

Figure 2: Plots of main effects (top) and two-factor interactions (bottom) for "err_40"

Valuable information is conveyed by the interaction effects. Noisy signals are better

captured by allocating more points to the initial LHS plan before starting the

adaptive procedure, while a few LHS points are enough for smooth signals;

however, starting with four points only, (the minimum number for allowing the

163


estimation of kriging parameters , , p and 2Z

), generally lowers the performance

especially when the criteria devoted to maximize the objective function (MaxF,

SR1) are adopted, see the n-Criterion interaction in Figure 2 and Figure 3 (left). The

same interaction also shows that the MaxF criterion is the most effective in

estimating the true error with n ranging in a quite wide interval, say from 20 to 34,

see Figure 3 (right).

Figure 3: geometric error (%) estimated by all criteria for signal 1, starting with 4

(left) and 32 (right) initial LHS points.

References:

[1] Krige DG. A statistical approach to some basic mine valuation problems on the

Witwatersrand. Journal of the Chemical, Metallurgical and Mining Society of South

Africa, 1951;52(6):119–39.

[2] Cressie NAC. Statistics for spatial data. 1st Edition New York: Wiley

Interscience; 1993.

[3] Pedone P., Vicario G., Romano D. Kriging-based sequential inspection plans for

coordinate measuring machines. Applied Stochastic Models in Business and

Industry 2009;25(2):133–49.

[4] Ascione R., Moroni G., Petrò S., Romano D. Adaptive inspection in coordinate

metrology based on kriging models, Precision Engineering, 2013;37(2013): 44–60.

164


165


Concept for a Miniaturized Machine-Tool-Module for the

Manufacturing of Micro-Components Operated at its

Resonance Frequency

C. Oberländer1, J.P. Wulfsberg1 1Helmut-Schmidt-University, University of the Federal Armed Forces Hamburg,

Germany

[email protected]

Abstract

The operation of conventional machine tools at resonance frequency is generally

avoided. High amplitudes generated by the mechanical resonance cause surface errors

on the workpiece and the machine tool can be damaged. This paper presents a new

concept for a machine-tool-module (MTM) which is operated at its resonance

frequency. It consists of a piezo actuator, a displacement amplifier and the tool itself.

By excitation the displacement amplifier at resonant frequency, very large amplitudes

at the tool can be achieved. The emphasis of this paper lies on the analysis of the

dynamic behavior of the displacement amplifier at its resonance frequency. The

results are compared with the static operation of the amplifier.

1 Introduction

The use of small machine-tool-modules (MTM) enables the application of new

technologies and functional principles, which are not suitable for the use in larger

machine tools. The basic idea behind the presented concept is to operate the MTM at

its resonance frequency, thus large amplitudes can be achieved to move the tool in

z-direction. The feed movement of the workpiece in x- and y-direction is realized by

the feed unit based on flexure systems, which has already been presented at

euspen [1] within the framework of Square-Foot-Manufacturing [2]. The MTM can

be used for manufacturing processes such as micro-cutting of thin foil or sheet,

micro-structuring of surfaces or minting. Unlike ultrasonic superimposed

manufacturing processes (e.g. ultrasonic cutting or ultrasonic stamping), this machine

concept only uses the amplified oscillation to move the tool in z-direction. There are

no additional actuators needed to move the tool in z-direction. Moreover, an

167




energetically favorable state predominates because of operating the structure at its

resonance range. Energy is withdrawn from the system only by damping and the

manufacturing process.

2 Design Concept

The machine-tool-module consists of an actuator for generating an oscillation, a

structure for transmitting and amplifying the oscillation (displacement amplifier) and

the tool itself.

ActuatorMechanical

AmplificationTool

Figure 1: Concept of the Machine-Tool-Module

The oscillation is generated by a piezoelectric actuator. The amplitude of the

oscillation generated by these kinds of actuators is typically in the range of 0-100µm

at frequencies up to 20,000Hz. To make this kind of oscillation usable for tool

movement, it must be transmitted and amplified.

The oscillation is transmitted and amplified by applying a transmission structure

based on flexure hinges which is specially adapted to the respective manufacturing

process. Flexure hinges, free of play, friction and wear, guarantee high dynamics

combined with high accuracy. As a first approach a geometry inspired by [3] is used

(Figure 2). Amplification is achieved in two stages: A pair of simple levers as the

first stage, a flexural bridge as the final stage. By the combination of lever and frame

solutions a highly stiff design with a rapid response can be achieved [3].

Figure 2: Displacement Amplifier

168


Figure 3: 7th mode shape at

1298 Hz

The mode shape of the transmission structure is used directly for moving the tool. By

exciting the structure in its resonance range, maximum amplitudes are generated at

the attachment point for the tool.

3 Simulation of Dynamic Behavior

3.1 Modal Analysis

As a condition for further analysis a modal

analysis of the amplifier geometry is performed.

The first 20 eigen modes and the their respectiv

mode shapes are calculated. Figure 3 shows the

7th mode shape at 1298Hz. This mode shape

was chosen for further analysis. The attachment

point for the tool oscillates with maximum

amplitude in z-direction and the amplifier only

oscillates in the x-y plane. For all simulations,

the FEM software ABAQUS was used.

3.2 Static and Dynamic Analysis

The actuator expansion ΔX is simulated by a displacement boundary condition of the

two contact surfaces between the actuator and the amplifier (ΔX = ΔX1 + ΔX2). The

output displacement ΔZ is measured at the attachment point for the tool (Figure 2).

The ratio of these two displacements is the displacement gain (ΔZ/ΔX). Four

different input displacements ΔX at three frequencies were analyzed. Table 1 shows

the results of the analysis. The mean displacement gain in static mode is about 1.18,

in dynamic mode at 150Hz about 2.04 and at resonance (1298Hz) 6.55 (Table1).

Table1: Displacement Gain in Static Mode and Dynamic Mode

Static Mode Dynamic Mode

150 Hz

Dynamic Mode

1298 Hz (Resonance)

Input ΔX

(µm)

Output ΔZ

(µm)

Displacement

Gain

Output ΔZ

(µm)

Displacement

Gain

Output ΔZ

(µm)

Displacement

Gain

10 11.7 1.17 20.4 2.04 66.1 6.61

20 23.6 1.18 40.9 2.05 128.6 6.43

30 35.3 1.18 59.7 1.99 194.7 6.49

50 58.9 1.18 103.5 2.07 334.5 6.69

169


Figure 4: Displacement ΔZ, ΔX1 and ΔX2 at 150Hz/ 1298Hz (Resonance)

By excitation of the amplifier at resonance frequency a three times larger

displacement ratio can be achieved. Figure 4 shows the displacement-time-graph for

ΔX = 30µm at 150Hz and 1298Hz.

4 Conclusion and Outlook

The simulation has shown that the excitation of the amplifier in the resonance range

leads to much larger amplitudes and displacement ratios. To make a statement on the

technical feasibility of the concept further investigations are necessary. For the

further development the influence of the attached tool and the manufacturing process

must be considered and an appropriate control system has to be designed. The size

and the influence of unwanted oscillation amplitudes in the X and Y directions must

be examined and reduced. In addition, the resulting stresses must be determined and

minimized by a geometry optimization process in order to maximize service life.

References:

[1] Kong, N., Grimske, S., Röhlig, B., Wulfsberg, J. P.: Flexure Based Feed Unit

for Long Feed Ranges: Concept and Design In: Proceedings of the 12th euspen

International Conference, Stockholm, June 2012

[2] Wulfsberg, J.P. , Kohrs, P., Grimske, S.; Röhlig, B.: Square Foot

Manufacturing - A new approach for desktop-sized reconfigurable machine

tools, Future Trends in Production Engineering - Proceedings of the WGP-

Conference, Berlin, Germany, 8th-9th June 2011; Publisher: Neugebauer, R.;

Schuh, G.; Uhlmann, E., 2012, Berlin

[3] Pozzi, M., King, T.: Piezoelectric Actuators in Micropositioning, Engineering

Science and Education Journal 10 (1), pp. 31-36, 2001

170


Concrete Based Parts for High Precision Applications

C. Hahm1, R. Theska1, K. John1, A. Flohr2, A. Dimmig-Osburg2 1Technische Universität Ilmenau, Germany 2Bauhaus-Universität Weimar, Germany

[email protected]

Abstract

In previous studies we have shown that Self Compacting Concrete (SCC) is a

promising alternative material for machine parts in high precision applications

conventionally designed of natural stone. Parts with comparable functional surface

finish and mechanical properties to those made of natural stone can be done in shorter

time at lower cost starting from small lot sizes. The developed “ready-to-use” primary

shaping process offers vast freedom of design compared to machined natural stone

[1]. In current studies, both moulding and post moulding processes have been

optimised. This article shows that a major improvement in long-term form stability,

time to stabilisation and surface roughness of moulded parts has been achieved.

1 Introduction

In previous studies the feasibility and the technology for achieving high precision

smooth and levelled functional surfaces at spacious parts with a flatness in the

micrometre range with standard SCC in a mould process have been demonstrated [1].

In current research the SCC mixtures were modified to achieve optimal material

properties comparable to parts made of natural stone. The modification can be done

in three ways: Using a high powder content (powder type), a viscosity modifying

agent (viscosity modifying agent type) or both (combination type) [2]. The latest

developed concretes are powder type SCCs. Two different cements (CEM I 42,5 R

and CEM II/A-LL 42,5 R), silica fume and fly ash as powder components were used.

In order to achieve high values in strength and Young´s modulus, basalt gravel and

sand were used for the coarse and fine aggregates respectively. The grading curve

was calculated according to Hüsken and Browers [3] to achieve the highest possible

packing density. A PCE superplasticiser ensures the viscosity of the SCC mixtures.

As a mould a reinforced frame design using plastic surfaces was fixed on top of a

171



high precision natural stone plate forming the reference face for the functional surface

of the part to mould. Polyethylene foil was applied as a barrier layer to protect the

natural stone and guarantee a surface roughness of less than 2 µm.


Concrete properties were tested in fresh and cured state and optimised for the

intended applications which demand higher strength and stiffness values as seen in

table 1. For the experiments concrete beams of 1400 x 80 x 160 mm³ were produced

and their roughness, flatness deviations and time dependent deformations were

measured, using a roughness meter and an autocollimator respectively.

Table1: Cured SCC properties compared with granite and standard concrete [1]

unit granite concrete SCC I SCC II

compr. strength [N/mm²] 250 - 360 5 - 55 110.1 109.2

flexure strength [N/mm²] 10 - 35 2 - 8 8.1 7.7

Young´s modulus [kN/mm²] 60 - 95 21.8 - 34.3 45.4 44.4

density [g/cm³] 2.90 2.0 - 2.6 2.48 2.47

2.1 Short range quality of moulded surfaces

Figure 1 displays the unfiltered roughness profiles of the samples. Beam 1 shows a

surface roughness Rz of 1.47 µm and a roughness average Ra of 0.16 µm without

post processing. These values are better than those of common granite surfaces used

Figure 1: roughness of special SCC surfaces

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for air bearings (Rz = 5.2 µm, Ra = 0.56 µm) [1] In comparison to earlier studies with

standard SCC the roughness has been decreased by 30%. The plot of beam 1,

moulded on a 100 µm thick foil shows a lower frequency but higher amplitude of

the waviness than beam 2 moulded on a 25 µm foil (figure 2). The visible waviness

is caused by inhomogeneities of the barrier foil’s stiffness and thickness. The search

for high quality foil that meets all requirements is part of future studies.

Figure 2: waviness of special concrete surfaces (left: beam 1, right: beam 2)

2.2 Long range quality of moulded surfaces

Quelling and shrinking has a significant influence on the flatness of concrete parts.

That is why concrete parts casted on a best flat standard plane bulge out in form of a

bending line. Figure 3 shows the maximum deformation of 1400 mm long sample

beams during a time period of over 100 days after casting. A granite reference beam

Figure 3: Long-term behaviour of the flatness of beams made with special SCC

173


is plotted for comparison. The maximum deformation of the test beam made of

SCC I is less than 60 µm which is an improvement of over 55% compared to

previous results [1].

After a period of about 4 weeks under laboratory conditions the concrete beams show

a good long-term stability comparable to the granite reference beam. To investigate

the influence of humidity, the SCC II and the granite reference beam were treated

with 100% relative humidity for two days between measurement day 10 and 11. The

deviation of the SCC II beam decreased by 25 µm while the granite beam did not

show any reaction to this treatment. Future studies will address this effect.

3 Conclusions

Concrete parts having functional surfaces with a roughness appropriate for aerostatic

guideways can be created by a “ready-to-use” mould process. Latest concrete

compositions show excellent long-term stability and mechanical properties compa-

rable to natural stone. Test beams with a length of 1400 mm and an absolute

maximum straightness deviation of 60 µm were casted. After the concrete has cured,

the long-term stability resembles the behaviour of granite. The research is now

focussing on the sensitivity to humidity and other environmental influences to create

parts that are feasible at normal environmental surroundings beyond the laboratory.

Acknowledgments:

The authors thank the German Federal Ministry of Economics and Technology for

the funding of this project.

References:

[1] Marius Berg, René Bernau, Torsten Erbe, Kay Bode, René Theska:

EUSPEN 2010 - Primary shaping of smooth and level guideway planes for high

precision applications

[2] Okamura, H.; Ouchi, M.; Skarendahl, A.; Petersson, Ö.: in Proceedings of 1st

int. Rilem symp. On SCC, Bagneux: RILEM Publications SARL; 1999, p. 3-14

[3] Hüsken, G.; Brouwers, H.J.H.: A new mix design concept for earth-moist

concrete: A theoretical and experimental study. Cement and Concrete Research 38

(2008), S. 1246-1259. + Erratum (2009)

174


Fast Nanometer Positioning System by Combining Fast

Resonant Mode and Accurate Piezostack Direct Drive

A. Santoso, J. Peirs, F. Al-Bender, D. Reynaerts

KU Leuven, Department of Mechanical Engineering, Belgium

[email protected]

Abstract

This work aims at the development of a fast nanometer positioning system,

combining the high speed capability of an ultrasonic piezomotors (resonant mode)

with the fine positioning capability of a piezostacks (direct-drive mode). The two

modes can be operated simultaneously with capability of achieving speed of more

than 200 mm/s and positioning accuracy of 10 nm.

1 Principle of Multi Drive Motor

The multi drive motor is able to do two different actuation modes simultaneously.

The first actuation mode is the resonant mode where the two piezos of the motor are

excited by two sine voltages with varying phase or amplitude. This excitation results

in an eliptical motion of the contact point. Since this contact point is preloaded

against a slider, it creates a stick and slip operational regime that results in

macroscopic drifting of the slider position. The second mode, the direct-drive mode,

drives the piezostack with a quasistatic voltage which results in microscopic

displacement with nanometer accuracy, over an operation stroke of ± 5 µm. The

principle of the two modes are shown in figure 1.

Figure 1: (Left) Principle of resonant mode; (Middle) Principle of direct-drive mode;

(Right) Picture of multi-drive motor mounted against rotational stage.

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2 Resonant Mode

In resonant mode, two different output control parameters are investigated. The first

is by adjusting the phase difference between the two sine. The response of the phase

input to output speed is shown in the graph below. The second input parameter is the

voltage amplitude. This type of control posseses nonlinear behavior in the form of a

deadzone. To acquire the advantage of each mode, phase control is implemented for

low velocity and gradually move to amplitude control when high velocity is desired.

Figure 2: (Left) phase regimes employed on the resonant operation; (Right) voltage

regimes employed on the resonant operation.

3 Direct-Drive Mode

The direct-drive utilizes two piezostacks driven by two independent quasistatic

voltages. To compensate the hysteresis of the piezostack, a maxwell slip hysteresis

compensation is implemented. In this research, two different actuation techniques are

investigated. The first technique drives one piezostack for moving the contact point to

one direction and drives the other piezostack for the oposite direction. The second

technique is implemented by giving a fix offset voltage and add an antagonistic

driving voltage to the two piezos. The two modes has been tested with good results,

with the first technique offers the advantage that for zero/initial position it requires

negligible voltage input, resulting in lower power consumption. Figure 3 shows the

comparison between the measured voltage input-output and the given input-

compensated output for the first method. The negative voltage shown on the X axis of

the Figure 3 (Left) means that the right piezo is given a positive voltage, while

positive voltage means that the left piezo is given a positive voltage.

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Figure 3: (Left) voltage input with its measured output position relation; (Right) input

and measured output position relation with Maxwell Slip hysteresis compensation.

4 Simultaneous Control of the two modes

A control scheme combining the two modes is implemented in a D-Space controller.

The global scheme of the control system is shown in figure 4. A more detailed

explanation of the scheme and its initial results of the applied scheme can be found

in [1].

Figure 4: Scheme of the simultaneous control for the two modes [1].

5 Experimental Results

For identification and testing the capabilities of the multi drive motor, a rotational

stage is implemented using ball bearing guides. The motor contact point is preloaded

against the ceramic contact ring fixed on the stage. The position is measured by

converting the angle acquired from a Renishaw Tonic rotary encoder. The converted

resolution of the sensor is 1.72 nm with a noise level of ± 6 nm and a maximum

allowable velocity of 50 mm/s.

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In figure 5 and table 1, the experimental results for optimized resonant mode are

shown. The tests involve a smoothed step position trajectory with specified

maximum velocity. The maximum jerk of this trajectory is limited to 5 mm/s3. The

results’ steady state error and the low velocity tracking error are at the same order of

magnitude as the sensor noise.

Figure 5: (Left) position input for constant velocity of 10 mm/s; (Right) the trajectory

following error

Table1: Error for different velocity value

Velocity Trajectory Length Max Error RMS Error

40 mm/s 400 mm 1.7 µm 0.147 µm

20 mm/s 200 mm 0.93 µm 0.062 µm

10 mm/s 100 mm 0.31 µm 0.046 µm

100 µm/s 1 mm 0.30 µm 0.033 µm

10 µm/s 0.1 mm 0.066 µm 0.007 µm

1 µm/s 0.01 mm 0.022 µm 0.004 µm

100 nm/s 0.001 mm 0.015 µm 0.003 µm

References:

[1] A. Santoso, J. Peirs, T. Janssens, and D. Reynaerts, Simultaneous Resonant and

Direct-Drive Control of a Piezomotor, for Combining Fast and Accurate Motion,

13th International Conference on New Actuators (2012), 730-733.

178


Towards the realization of the new INRIM angle comparator

M. Pisani and M. Astrua

Istituto Nazionale di Ricerca Metrologica, INRIM, Italy

[email protected]

Abstract

A new angle comparator is under construction at INRIM. The goal uncertainty is 0.01

arcsec (50 nrad). The device will be based on a double pneumostatic air bearing and

will exploit the rotating encoder principle. A prototype has been built to demonstrate

the effectiveness of the principle and to test electronics and software. The prototype

and preliminary results are presented as well as the design of the comparator.

1 Introduction

Angle measurements is one of the critical issues in precision mechanics metrology.

All modern angle measurement instruments are based on state of the art angle

encoders. Technological progress achieved in the last decades has allowed

tremendous improvement in the encoder performances. Resolutions down to 0.01’’

and accuracy better than 1’’ are commonly achieved. The calibration of such divided

circles is a basic task of angle metrology. Hereby, calibration means the

determination of the division errors as deviations from nominal circular division. The

main error sources of angle encoders are the non-uniformity of the grating spacing

(due to manufacturing errors or misalignment) and the nonlinearity occurring when

subdividing the grating pitch in small parts (fringe interpolation error).

The realization of rotary tables (RT) having extremely high accuracy is the preferred

solution for this kind of calibration. Main national metrology institutes have designed

and realized their own unique instrument based on different technological solutions to

achieve this goal [1,2].

Until now the encoder calibration facility at INRIM was based on precision index

tables that, although having excellent accuracy, require extremely long fully manual

procedure. INRIM has recently afforded the realization of a novel high precision

automated RT. A preliminary demonstrator has been built and described in the next

section.

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2 The prototype

The measurement principle is sketched in figure 1. The instrument is based on a

principle pioneered by E. W. Palmer in 1984 [3]. A pair of continuously rotating

encoders read by two pairs of heads, one fixed with respect to the laboratory frame

and a second rotating with the measurement drum. The angle measurement is based

on the phase difference between the fixed head signal (used as a reference) and the

rotating head. The phase measurement is intrinsically free from nonlinearities and the

encoder errors are cancelled by the average made each complete revolution of the

encoder.

Figure 1: Schematic of the rotating encoder principle. D continuously rotating optical

encoder; F: reading head fixed to the reference frame BA; E: reading head fixed to

the measuring table A; G: phase measurement.

2.1 Mechanical structure

The demonstrator is based on a Heidenhain ring encoder (model ERA 4200, 40000

lines with 20 μm spacing), mounted on a precision air bearing (Precision Instrument

Inc.), driven through a belt by a microstep motor (Oriental Motors). Two heads are

faced to the ring about 180° one respect to the other. One is fixed to the table and

represent the reference head of the system. The second is mounted on a piezo-

capacitive transducer (PI P-753) capable of 25 μm displacement, representing the

moving head of the comparator. A controlled movement of the piezo actuator

simulates a movement of the comparator corresponding to an angular rotation of 1.62

arcsec each micrometer of the actuator.

Each head generates two quadrature sinusoidal signals. Said signals are amplified and

sent to an analog to digital converter board (ADC, NI‐USB‐6259 BNC).

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Figure 2: A) picture of the prototype. In the center the rotating ring encoder mounted

on the pneumostatic spindle. Upper right corner the microstep motor. Right the fixed

head. Left the moving head. B) detail of the moving head mounted on the piezo-

capacitive actuator capable of simulating microradiant rotations.

2.2 Electronics and software

The purpose of the electronics is the conditioning of the signals generated by the

heads and the generation of a trigger signal for the ADC. The differential signals

coming from the heads are amplified by low noise and fast differential amplifiers

built to the purpose. The conditioned signals are sent to the ADC board having 16 bit

resolution and a maximum sample rate of 1.25 MS/s. The encoder is rotated at 90°/s

so, the base head signal has about 10 kHz frequency. The signals are sampled at 8

points per cycle (80 kHz). In order to avoid spurious noise coming from the beat of

the sampling frequency and the signal, we decided for a synchronous sampling of the

reference signal. A Phase Locked Loop (PLL) circuit has been built to the purpose.

The software (based on LabView®) elaborates the signals captured with the ADC

boards and triggered with the phase locked clock according the following simplified

steps. The reference signal is mixed (multiplied) with the sine and the cosine signals

of the measurement head. The two quadrature signals now represent the phase vector

which carries the angular information. Each complete phase revolution corresponds to

a shift of one encoder line (32.4 arcsec). A Matlab® based algorithm calculates the

instantaneous phase angle. The phase is than averaged over the entire revolution of

the encoder. A counting logic measures the integer part of the phase (the number of

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complete phase revolutions) and adds or subtract the integer to the fractional

measurement. The final value is converted in arcseconds and the result is stored.

3 Results

We have performed two tests. The first is to check for the long term stability of the

device. In figure 5 two typical long term acquisition runs are captured. The long term

stability is around 0.01 arcsec per hour. That corresponds to a mechanical drift of the

two reading heads better than 10 nm per hour, compliant with the expected thermal

drift of the overall structure.

Figure 5: typical drift over 3-4 hours period. The two curves are respectively with the

piezo-actuator switched off and switched on.

In the second test we have driven the piezo-capacitive actuator with a square signal

having 100 s period. The step height is 12 nm corresponding to a 0.02 arcsec peak to

peak rotation. In figure 6 ten consecutive 400 s records are plotted. The vertical

dispersion is compliant with the above reported 0.01 arcsec per hour drift.

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Figure 6: response to a 0.02 arcsec p.p. step. Full acquisition time is 4000 s.

5 Conclusions

A demonstrator for the development of the electronics and the software of the new

INRIM angle comparator have been presented. The preliminary results are compliant

to the expected accuracy of the system. On the basis of the results above the angle

comparator, which will be based on a double air bearing structure, will be designed

and built.

References:

[1] R Probst, R Wittekopf, M Krause, H Dangschat and A Ernst, The new PTB

angle comparator, 1998 Meas. Sci. Technol. 9 1059

[2] Jack A. Stone Jr, M Amer, Bryon S. Faust, Jay H. Zimmerman, Angle Metrology

Using AAMACS and Two Small-Angle Measurement Systems 2003, Proceedings

of SPIE 5190 pp. 146 - 155

[3] E.W. Palmer, Goniometer with continuously rotating gratings for use as an

angle standard, Precision Engineering 6359(88)90033-5

183


Geometrical-based approach for flexure mechanism design

T.J. Teo1, G. Z. Lum1,2,3, G.L. Yang1, S. H. Yeo2, M. Sitti3

1Singapore Institute of Manufacturing Technology, Singapore 2Nanyang Technological University, Singapore 3 Carnegie Mellon University, United States.

[email protected]

Abstract

This paper introduces an alternate approach for designing a flexure-based parallel

mechanism (FPM). It involves a systematic design methodology that couples classical

kinematics with modern geometrical optimization techniques. At sub-chain level, a

novel topological and structural optimization technique is introduced to synthesize

and optimize the geometry of the joint/limb based on desired stiffness characteristics.

At configuration level, the moving masses and stiffness of the entire FPM are

optimized based on desired dynamics. Using this new design approach, the system

characteristics of the FPM is optimal and deterministic. This paper presents how this

geometrical-based approach was used to design a 3-axes planar motion FPM.

1 Introduction

For many years, an exact constraint method is a well-established kinematic approach

for designing any flexure-based joint/mechanism [1-2]. Even approaches introduced

lately, e.g., the Freedom and Constraint Topology (FACT) [3] and those derived from

screw theory [4] etc., are variants of the exact constraint method. These approaches

have their merits when the constraints are ideal and the size of the synthesized

mechanism is unlimited. Yet, they could only synthesis the topology of a mechanism

based on ideal constraint conditions rather than delivering an optimal design based on

desired system dynamics. This paper presents systematic design methodology, which

couples classical kinematics with modern geometrical optimization techniques, to

design an optimal FPM based on desired stiffness characteristics and moving masses.

2 Systematic design methodology

The first step of the design methodology is to synthesize the type of parallel-

kinematic configuration based on the desired degrees-of-freedom (DOF) and task etc.

At sub-chain level, a novel topological optimization is proposed to synthesize the

184



flexure-based joints/limbs of the selected configuration. Concurrently, the structure

optimization delivers the optimal geometries based on given stiffness characteristics.

At configuration level, the moving mass of each joint/limb is predicted through mass

condensation method. Hence, both stiffness and moving masses of the entire FPM

can be optimized based on the desired workspace and size constraint. In this paper,

design of a planar motion FPM is used to demonstrate this proposed methodology.

Figure 1: Block diagram representation of the systematic design methodology.

2.1 Mechanism synthesis

To achieve a 3-DOF planar motion, i.e., X-Y-z, 3RRR, 3PRR, and 3PPR [5] are

possible parallel-kinematics configurations (prismatic; P and revolute; R). 3PPR was

chosen as the compliant P joints are more deterministic than the compliant R joints.

The schematic of 3PPR is shown in Fig. 2 where the moving platform is connected to

the fixed base by three identical parallel chains. Each chain comprises of a serially-

connected active P joint and a passive RP joint. The overall size is 300x300mm2 to

amplify the errors for proper evaluation while the workspace is targeted at 4mm2 x 2°.

Figure 2: Schematic representation of 3PPR and its stiffness modelling.

2.2 Topological optimization: Mechanism-based approach

In this work, a hybrid topological and structural optimization technique is used to

deliver optimal designs for the joints. Termed as mechanism-based approach,

elementary kinematics chains are used as basic genes for the joint optimization,

which runs on Generic Algorithm. Here, a generic 4-bar kinematics chain is used to

synthesize the 1-DOF P joint and a generic 5-bar kinematics chain is used for 2-DOF

185


PR joint synthesis. For P joint, C11 needs to be very compliance compared to other

components of the compliance matrix. Hence, the objective function is formulated as

subjected to

(1)

For PR joint, both C11 and C66 need to be very compliance compared to other

components of the compliance matrix. Thus, the objective function is formulated as

subjected to

(2)

Subsequently, optimizations are conducted based on these objective functions. The

evolutions from the basic genes (kinematic chains) to optimal joint designs in both

topological and structural forms are shown in Fig. 3.

Figure 3: Concurrent evolution of topology and structure for both P and PR joints.

2.3 Configuration level: Overall mass and stiffness optimization

At configuration level, optimization constrains are based on desired workspace and

size constraint. Using the compliance matrices derived from the optimal joints, the

stiffness of the FPM can be obtained through kinematic stiffness modelling (Fig. 2).

Figure 4: Others optimization parameters Figure 5: Mass prediction.

At this stage, only the stiffness of the proposed FPM was optimized with the aim of

maximizing all non-actuating stiffness while minimizing all actuating stiffness

through parameters such as the flexure length in P joint and the base of the PR joint

(Fig. 4). In future, the proposed dynamics optimization needs to be done concurrently

with a new mass optimization algorithm; using mass condensation technique, which

186


at this stage has proven to have good prediction on the first 3 frequency modes of

optimal P joint design as shown in Fig. 5.


Stiffness evaluation is conducted using PICO-motor to induce load while the motion

of the end-effector is captured by a non-contact 3D-GOM system. At this stage of

research, only the actuating stiffness of the developed FPM are evaluated. Results

plotted in Fig. 6 show that the actuating compliance in Y-axis and about Z-axis are

3.91×10-5 m/N and 0.0156 rad/Nm respectively. Comparing with the theoretical

prediction of 3.39 ×10-5 m/N and 0.0133 rad/Nm, such small deviations prove that

this approach is good for designing FPM based on desired stiffness. The prototype

has achieved positioning and angular resolutions of 50nm and 0.2 arcsec respectively

throughout a workspace of 4mm2 x 2°.

Figure 6: Developed prototype and measured compliance in Y-axis and about Z-axis.

4 Conclusion

This paper presented a geometrical-based approach to design a FPM. A novel hybrid

topology optimization technique for stiffness optimization, and a new technique for

mass optimization are introduced. Experimental results show that this approach is

good for designing FPM based on desired system dynamics, i.e., stiffness and mass.

References:

[1] AH. Slocum, Precision Machine Design, Prentice-Hall, Inc.; 1992.

[2] DL. Blanding, Principles of Exact Constraint Mechanical Design, Kodak; 1992.

[3] JB. Hopkins, ML. Culpepper, Synthesis of precision serial flexure systems using

freedom and constraint topologies (FACT), Prec. Eng., 2011 (35), 638 – 649.

[4] JB. Hopkins, RM. Panas, Design of flexure-based precision transmission

mechanism using screw theory, Prec. Eng., 2013 (37), 299 – 307.

[5] GL. Yang, W. Lin, TJ. Teo, CM Kiew, "A flexure-based planar parallel

nanopositioner with partially decoupled kinematic architecture," Proc. of

EUSPEN2008, 18 – 22 May, Zurich, Switzerland, 2008, vol. 1, 160 – 165.

187


A New Approach on Reducing Thermal Impacts on High

Precision Machine Tools

M. Fritz1, Dr. D. Janitza1 1KERN Microtechnik GmbH, Germany

[email protected]

Abstract

This paper deals with a new thermal concept for high precision milling machine tools,

that has been designed and verified at the KERN Microtechnik GmbH and is now

part of the latest KERN machine tool series. By combining aluminum light weight

construction with a high precision temperature management it was possible to

overcome a traditional conflict of goals between machine dynamics and its

temperature stability.

1 Introduction

In order to machine high precision parts within their tolerances it is necessary to cope

with a multitude of different influences and their resulting errors. Bryan [1] and Weck

[2] claim, that up to 70% of today’s errors on machined parts are due to thermal

effects within in the machine tool, the spindle, the work pieces, the cooling liquid,

etc.

2 State of the art

Dealing with these effects, traditional guide lines for designing thermal stable

machine tools have been established. Trying to configure the whole machine system

as a thermal low pass filter, machine designers have always aimed at designing a

system with a very low response to changes in the environmental conditions.

Therefore materials with a very high thermal capacity and very low thermal

conductivity have been chosen (e.g. heavy weighted iron casts, polymer concrete) not

only for the machine basis but also for the axes and other moving parts. On the one

hand theses designs are insensitive to short thermal disturbances and high frequency

vibrations. On the other hand they have quite a high mass, resulting in long time

188


thermal adaption cycles, the necessity of long term warm up phases and huge

temperature gradients, that are very complex in terms of modeling, simulation and

compensation. Research in the last 20 years has been focused on this problem [2, 3],

generally resulting in compensations that are based on highly complex models with a

large number of input variables.

2 A new concept

In our work a different approach is presented, that more or less reverses the design

principles mentioned above. Designing “fast” thermal assemblies with low thermal

capacity and high thermal conductivity leads to lightweight constructions with a very

short thermal response time. The resulting, dynamic system is characterized by short

warm up cycles and the avoidance of temperature gradients within the assembly.

2.1 Avoidance of temperature gradients – homogenous temperature

Thermal displacements in machines tools due to temperature gradients are much

higher than due to homogenous warming. These effects are even worse on slim

components such as spindle sleeves or portal frame bridges. Asymmetrical thermal

loads on these parts lead to massive mechanical displacements being caused mainly

by the emerging temperature gradients.

Therefore, the obvious first step towards thermal stable constructions is the avoidance

of temperature gradients within the construction elements by choosing materials with

a high thermal conductivity (e.g. aluminum alloys). Large cross sections within the

affected components in combination with short distances to the next heat sink

guarantee a fast heat transfer. Following these design principals, leads to components

that, even when thermally asymmetrical loaded, tend to have a very homogenous

temperature distribution and therefore a very predictable homogenous thermal

displacement.

2.2 Keeping the homogenous temperature constant

In order to guarantee a stable machine tool behavior it is essential to keep the

homogenous temperature as constant as possible. Conventional approaches therefore

189


use components with high weight densities and materials with high thermal

capacities. The biggest advantage of these designs is a very slow reaction due to

thermal disturbances. The biggest disadvantage of these designs is the high mass of

the components and the necessity of long warming up cycles.

The presented concept is based on the idea of using low weight density components

with a low thermal capacity. By installing an accurate temperature management

system the thermal capacity is virtually raised. Using the temperature management

system to systematically deprive the heat out of the components the design works like

it has got an infinite thermal capacity. That means a change in the environment

conditions, only results in a negligible temperature change within the component. The

temperature of the component is more or less completely controlled by the

temperature management system and can be changed instantly. Therefore warm up

cycles can be reduced to a minimum. Furthermore, this approach offers the possibility

to use low weight density materials, leading to a lot of advantages such as less energy

consumption and better dynamic behavior.

2.3 Dissipate the heat where it origins

The main heat sources within a machine tool are the working spindle, the drives,

process heat and heat due to friction within the guides and bearings. Heat from these

sources is unavoidable most of the times. In order to evade temperature gradients it is

necessary to dissipate this heat as close to its origin as possible resulting in another

big advantage of reducing the warm up cycles. Placing the heat sinks in proximity to

the heat sources is the foundation of an accurate temperature management with fast

and stable feedback control cycles (see chapter 2.4).

2.4 Temperature management

The temperature management system has specifically been adapted to the machine

design and is based on the following principals:

Temperature of the coolant fluid must be constant even on changing

environment conditions and machine conditions.

Flow rates must be high in order to ensure small temperature variations

between the inlets and the outlets of the cooling circuit.

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High cross sections lead to low back pressures and lower pump capacities

Heat will not dissipate into the tool shop but into central water chillers

Cooling and heating of the coolant lubricant in order to achieve

temperature controlled tools and work pieces

3 Machine design

The theoretical considerations as well as the results of practical experiments have

been transferred to the design of the latest KERN machine tool. Despite all traditional

design principles the axis components have been build in aluminum based light

weight designs. High thermal conductivity of aluminum leads to lower temperature

gradients. The increased wall thicknesses improve the vibration and thermal behavior

of the design. By reducing the components mass (approx. 20%), drive forces could be

reduced as well. This reduces energy consumption and the thermal loss of the direct

drives which lead to thermal improvement of the whole machine tool.

4 Conclusion

Combining the four approaches mentioned in chapter 2 within a new machine tool,

has created an exciting base for a long term stable production of high precision parts

with short warm up phases and a minimum of thermal based positioning errors. By

actually applying this concept to a serial machine tool KERN has created a

worldwide novelty in the design and functionality of industrial high precision

machine tools, that has been proven right by one year of experience (with several

machines) and the consistently positive feedback of the machine operators.

References:

[1] J. Bryan. International status of thermal error research. Annals of the CIRP,

1990.

[2] M.Weck, P. McKeown, and R. Bonse. Reduction and compensation of thermal

errors in machine tools. Annals of the CIRP, 1995.

[3] J. Mayr et. al., Thermal issues in Machine tools, Annals of the CIRP, 2012.

191


Long Range Precision Stage Using Multi Bar Mirrors

Siwoong Woo, Dahoon Ahn, Jaehyun Park, Daegab GweonKorea Advanced Institute of Science and Technology (KAIST), Republic of Korea

[email protected]

Abstract

In long range precision stage systems, laser interferometers are used to measure the

position of stage’s target mover. Long length mirrors called bar-mirrors have to be

used with laser interferometers to reflect laser. The length of Bar-mirror is

proportional to the range of precision stage systems. So, the long length bar-mirrors

are must for the long range precision stage systems. However, as the length of bar-

mirrors is lengthen, the flatness error of bar-mirror become large. In addition, long

length bar-mirror is hard to make, and expensive. Newly proposed long range

precision stage system is made up except long length bar-mirror.

Basic concept of the proposed system is using numerous short bar-mirrors instead of

one long length bar-mirror. There are two main problems to realize proposed system.

First, there is the alignment error. The alignment error includes the offset error and

the tilt error. The offset error means a linear misalignment, and the tilt error means an

angular misalignment of each bar-mirrors. To solve the alignment error problem,

measure the alignment error and compensate it.

Second, there are discrete parts between bar-mirrors. To reflect laser beam, bar-

mirrors do not have to discrete parts on mirror surface. In the proposed system, if the

discrete part overlapped with laser beam path, the feedback laser interferometer

signal is switched extra laser interferometer’s signal.

In this paper, propose new type precision stage using numerous short bar-mirrors, and

answers are given to solve two problems of new system. By evaluation experiment,

evaluate performences of proposed system.

1 Principle of alignment error measurement

There are many methods to separate these errors and to measure workpiece flatness

profiles. The inclination method [1] and the generalized two-point method [2] have

192


1

2

( ) ( ) ( )

( ) ( ) ( ) ( )

n n n

n n n n

d x f x t x

d x f x s t x s x

(1)

2 1

( ) ( ) ( )

( ) ( ) ( )

n n n

n n n

F x f x s f x

d x d x s x

(2)

'

2 1

( ) ( ) /

( ( ) ( ) ( )) /

n n

n n n

F x F x s

d x d x s x s

(3)

'

1

'1

1 2 1

( ) ( )

( ) ( )

( ) ( ( ) ( ) ( ))

n

n ii

n n

n n n n

p x F x s

p x F x s

p x d x d x s x s

(4)

Figure 1: Principle of the generalized two-point method

been proposed for that purpose. In these methods, two displacement probes are used

to measure the flatness profile. The straightness motion of the stage is canceled by

the differences of probes’ output, and the flatness profile of bar-mirror is obtained

by simple data processing operations. The combined method [3], which combines

the advantages of the inclination method and the generalized two-point method, is

proposed. Also, to realize high accuracy profile measurement many kinds of three-

point methods [4, 5] are proposed.

The generalized two-point method is selected to measure the flatness profile of bar-

mirrors. Figure 1 shows the principle of the generalized two-point method

schematically. Two displacement sensing probes are fixed and can measure the

flatness profile of bar-mirror while the stage moves. Assume that the flatness profile

of bar-mirror is f(x), the straightness motion of the stage is t(x), and the yaw motion

of the stage is θ(x).

2 Sensor switching method

The proposed system has four laser interferometers to measure x, y and theta-z

positions. . Generally in three degree of freedom systems, three laser interferometers

are used to measure x, y and theta-z positions. However, proposed system has the

discrete parts on mirror surface which are impossible to reflect laser beam. Figure 2 is

the proposed four sensor system’s schematic diagram. The sensor 1 & 2 measure x

and theta-z position. The sensor 3 & 4 are used for measuring y position by switching.

193


Figure 2: Schematic diagram of proposed system

If the discrete part overlapped with laser beam path, the feedback laser interferometer

signal is switched extra laser interferometer’s signal.

3 Experiment

By experiment, to confirm these proposed system. Figure 3 shows the proposed

system and experiment setup. Two bar-mirrors are aligned in x-axis on precision

stage’s mover. The length of x-axis bar-mirror is 150 mm and y-axis is 300 mm.

To measure alignment error, capacitive sensors are used. The capacitive sensor

interval is 15 mm and sampling period is 3 mm.

To evaluate experiment result, laser calibrator (Renishaw, ML10) is used.

Straightness errors are measured five times each. Figure 4 is the result of

straightness error measurement. In table 1, there are quantities of errors

compensation.

Figure 3: Photography of the precision stage used to experiment

194


Figure 4: Measurement data of the straightness errorTable 1: The quantites of the straightness error

Before Compensation After Compensation

Straightness Error 273.33 um 0.76 um

4 Conclusion

In this paper, new concept precision stage using multi bar-mirror was suggested.

Found the solutions to solve alignment error problem and discrete part between bar-

mirrors problem. And after evaluate precision stage’s straightness error using laser

calibrator.

References:

[1] Makosch G, Drollinger B. Surface profile measurement with a scanning

differential ac interferometer. Applied Optics. 1984; 23: 4544-4553.

[2] Omar B.A, Holloway A.J, Emmony D.C. Differential phase quadrature

surface profiling interferometer. Applied Optics. 1990; 29: 4715-4719.

[3] Gao W, Kiyono S. High accuracy profile measurement of a machined surface

by the combined method. Measurement. 1996; 29: 4715-4719

[4] Gao W, Kiyono S. On-machine roundness measurement of cylindrical

workpieces by the combined three-point method. Measurement. 1997; 21:

147-156

[5] Li CJ, Li S-Y, Yu J. High resolution error separation technique for in-situ

straightness measurement of machine tools and workpiece. Mechatronics.

1996; 6: 337-347

195


Feasibility study on a spindle supported by high stiffness

water hydrostatic bearings for ultra-precision machine tool

Yohichi Nakao 1, Kohei Yamada 1, Kotaro Wakabayashi 1, Kenji Suzuki 1 1 Kanagawa University, Japan

[email protected]

Abstract

A prototype of a spindle supported by water hydrostatic thrust bearings is considered

in the present paper. Stiffness of the bearing of the spindle is designed to be 1 kN/m.

Due to lack of the lubricative property of water, several materials for the bearing parts

are considered.

1 Introduction

A design of the spindle supported by water hydrostatic bearings for ultra-precision

machine tools is considered in this paper. Bearing stiffness for the ultra-precision

machine tools is a crucial characteristic to be taken into account in the spindle

design. Besides the bearing stiffness, precise rotational motion accuracy and thermal

stability of the spindle are important as well. In order to meet the requirements, the

hydrostatic bearings are in many cases used to support the spindle. Among them, the

water hydrostatic bearings[1]-[4] can be a suitable candidate for the bearing, because

of the low viscosity and high thermal conductivity of water.

In general, the stiffness of the hydrostatic bearings increases with the increase in the

supply pressure of the lubricant fluid. Thus, the stiffness of oil or water hydrostatic

bearings is relatively easy to increase. However, the higher viscosity of oil must be a

disadvantage in the higher spindle speed operation. On the other hand, in the case of

air bearings, allowable maximum pressure is restricted due to the compressibility of

air. These considerations indicate the water hydrostatic bearing is suitable for the

spindle application if the higher bearing stiffness is required.

Accordingly, this paper studies a spindle design supported by the water hydrostatic

bearings. An objective of the new spindle design is to aim the thrust bearing stiffness

of 1 kN/m. The paper thus considers a design of the water hydrostatic thrust

bearings. Specifically, we consider the influences of the supply pressure and bearing

gap on the achievable bearing stiffness.

196



In the design of the water hydrostatic bearings, the materials of the mating surfaces

should be appropriately chosen. We thus consider materials for the bearing pads to

made of gun metal, engineering ceramics, carbon impregnated resin and

polyetheretherketone (PEEK) as the candidates of the materials. Before designing

actual spindle for the ultra-precision machine tool, a spindle with simplified

structure is designed as a prototype. The structure and materials of the spindle is

presented with the characteristics of the water hydrostatic bearings of the spindle.

(a) Cross section of spindle (b) Exploded view

Figure 1: Structure of designed spindle with water hydrostatic thrust bearings

2 Designed prototype spindle with water hydrostatic thrust bearings

A structure of designed prototype spindle is given in Fig. 1. The spindle is equipped

with water hydrostatic thrust bearings. Meanwhile sliding journal bearings, instead of

water hydrostatic radial bearings, are used to support the spindle rotor in the radial

directions. Thus we consider the design of the water hydrostatic thrust bearings for

the spindle in this paper.

All the parts of the spindle except for the bearing pads and sliding journal bearings

are made of the stainless steel. As shown in Fig. 1, the rotor is placed between two

bearing pads that are fixed on side covers. The bearing pads shown in Fig. 2 are

exchangeable so that various pads made of different materials can be tested in

experiments. In the present study, several bearing pads are made of gun metal,

engineering ceramics, engineering plastics and carbon impregnated resin. Among the

materials, it is considered that the gun metal and the carbon impregnated resin have

Spindle rotor

Thrust bearing pad

Restrictor

Thrust bearing fixing ring

Spacer plate

Radial bearing

Side cover Key

Spacer plate Casing

Side cover

Restrictor

Thrust bearing

Bearing pad

Sliding bearing

Rotor

Drain port

Water supply

Water supply

197


better wear characteristics in the water lubrication. In

particular, the carbon impregnated resin has many industrial

applications in water lubrication, because of its self-

lubricating property. Operating tests are currently being

prepared using different pads to find suitable material for

the water hydrostatic bearings. The results will be compared

in the future work. The sliding journal bearings are made of the engineering plastics.

As well known, the bearing gap is a critical parameter determining the bearing

stiffness. Thus, the spindle is designed so that the gap of the thrust bearing can be

changed using a spacer plate that is placed between spindle casing and a side cover.

Bearing restrictors that are carefully designed[4] are inserted in the both side plates.

3 Design of water hydrostatic thrust bearing

For precision machine tool applications, the bearing has to be designed with careful

considerations on the bearing stiffness. Assuming cutting force is 1 N during single

point diamond turning. The resultant displacement of the bearing due to the cutting

force must be minimized. If the displacement is needed to be less than 1 nm, the

required bearing stiffness reaches 1 kN/m. For next generation of the precision

machining, a spindle with the stiffness of 1 kN/m is highly desired.

The thrust bearing of the designed spindle is a multi-recess opposed pad bearing. The

outer and inter diameters of the bearing pad are 82 mm and 32 mm, respectively. The

stiffness of the bearing is calculated as given in Fig. 3. It is given for various bearing

gaps h0, showing the stiffness of 1 kN/m is achieved if the gap and the supply

pressure are 17 m and 3 MPa, respectively. In order to prepare a water pump for the

designed hydrostatic bearings the required flow rate for the bearings is needed for

estimating the power and size of the pump. Therefore, the water flow rate is

calculated for various gaps and the supply pressures as shown in Fig. 4. This indicates

that the required water flow rate is about 7.5 L/min, thus the power of pump becomes

375 W.

In this spindle design, the spindle speed in the normal operation is considered to be

2,000 - 3,000 min-1. The loss of the power due to the viscosity of water during spindle

rotation must be taken into account. For instance, the supply of water flow increases

Figure 2: Bearing pad

198


the temperature of water. This must be minimized by an appropriate water

temperature control system that will be designed in our future work.

Figure 3: Bearing stiffness Figure 4: Flow rate for designed bearing

4 Summary

A spindle design supported by water hydrostatic thrust bearings was considered in the

paper. An objective bearing stiffness was 1 kN/m in order to improve machining

accuracy of the single point diamond turning. It is then verified that the bearing

stiffness of 1kN/m is obtained by the supply pressure 3 MPa and the gap of 17 m

for given bearing sizes; inner and outer diameters are 32 mm and 82 mm,

respectively. The bearing part of the spindle is exchangeable. Thus various materials

will be tested in the experimental works for finding suitable material combination for

the water hydrostatic bearings.

Acknowledgement

This research work is financially supported by the Mitutoyo Association for Science

and Technology.

References:

[1] Y. Nakao, M. Mimura, and F. Kobayashi, Water Energy Drive Spindle

Supported by Water Hydrostatic Bearing for Ultra-Precision Machine Tool, Proc. of

ASPE 2003 Annual Meeting, pp. 199-202, 2003,.

[2] A. Slocum, et al., Design of Self-Compensated, Water-Hydrostatic Bearings,

Precision Engineering, Vol. 17, No. 3, pp. 173-185, 1995.

[3] Y. Nakao, M. Kawakami, Design of Water Driven Stage, Proceedings of 9th

International Conference of the European Society for Precision Engineering and

Nanotechnology, Vol. 1, pp. 200-203, 2009.

[4] Y. Nakao, S. Nakatsugawa, M. Komori and K. Suzuki, Design of Short-Pipe

Restrictor of Hydrostatic Thrust Bearings, Proc. of ASME 2012 International

Mechanical Congress and Exposition, CD-ROM, 2012.

0 0.5 1 1.5 2 2.5 3 3.50

500

1000

1500

2000

2500

Supply pressure Ps [MPa]

Sti

ffn

ess

K [

N/

m]

h0=15m

h0=14m

h0=13m

h0=12m

h0=11m

h0=10m

h0=16m

h0=17m

h0=18m

h0=19m

h0=20m

0 0.5 1 1.5 2 2.5 3 3.50

2

4

6

8

10

12

14

16

Supply pressure Ps [MPa]

Flo

wra

te Q

[L

/min

]

h0=15m

h0=16m

h0=17m

h0=18m

h0=19m

h0=20m

h0=14m

h0=13m

h0=12m

h0=11m

h0=10m

199


Investigations of a small Machine Tool with CFRP-frame

Hoffmeister, H.-W.1, Gerdes, A.1, Verl, A.2, Wurst, K.-H.2, Heinze, T.2, Batke, C.2

1TU Braunschweig, Institute of Machine Tools and Production Technology,

Braunschweig, Germany 2Universität Stuttgart, Institute for Control Engineering of Machine Tools and

Manufacturing Units, Stuttgart, Germany

[email protected]

Introduction

Today microcomponents are one of the most important parts in industry as they are

used in optical products or in precision and medical applications. However, the

installation space of the machine tools used for manufacturing of small parts still is

too oversized compared to the necessary working space regarding the small

dimensions of the workpieces [1, 2]. Within the Priority Programme 1476 “Small

Machine Tools for Small Parts”, funded by the German Research Foundation (DFG),

a novel kinematic module based on cooperative and inverse motion was developed to

minimize the working space of the whole machine frame, as well as the moving

masses and the kinetic energy [3].

Workpiece-Axis

Machinespace 2L

Machine Frame

Carriage

Workpiece

L

Werkzeug

Workpiece-Axis

Machinespace 3/2 L

Carriage

Tool-Axis

Workpiece

Tool

Carriage

Stroke L ½ L

½ L

Stroke L

Standard Design

(STD) Cooperative Motion

(COOP)

Figure 1: Machine tool design with cooperative motion [3]

1 Energy consumption using Cooperative Motion

The cooperative kinematic reduces quantities like stroke, velocity and motor current.

Thus, the consumed energy of a cooperatively driven system differs from that of a

standard machine design. Energy is needed to drive both, the workpiece- and tool

carriage. The amount of energy depends on the inertia and friction of the carriage.

200


To provide the electric energy, servo amplifiers are needed for each drive. Another

aspect of the energy consumption is the loss due to friction forces occurring within

the drive system. A comparison of the friction losses of non-cooperative (STD)

motion and cooperative motion (COOP) is shown in Figure 2. It can be seen that if

the drive has linear friction characteristics and cooperative motion is used, the

frictional loss would be reduced by 50% due to the reduction of velocity and stroke.

In case of coulomb friction characteristics, cooperative motion causes the same

friction losses as in a non-cooperative setup. In case of a stribeck friction

characteristic the frictional losses can only be reduced if the resulting drive velocity

after splitting the motion profile still is above the stribeck velocity v0.

Figure 2: Influence of friction characteristics on the total friction power

Measurements on the prototype [3] showed that the total power consumption related

to the kinetic energy could be reduced to about 50% (Figure 3a). However,

measurements of the total power consumption of the electrical cabinet showed an

increase of about 55% in the cooperative mode (Figure 3b). The reason is the

necessary additional drive. It doubles the electrical losses within the amplifiers.

Energy savings with cooperative motion can only be achieved if the kinetic and

frictional power is higher than the electrical losses within the power supply.

0 10 20 30 40 50 60 70 80 90 100

KOOP

STD

Power consumption [W]

Actual components in the control cabinet:- Servo drive X1Y1 - Servo drive X2Y2- Servo drive Z1 - Power supply

0,0E+00 1,0E-04 2,0E-04 3,0E-04

KOOP

STD

Kinetic energy [J]

E_Kin Axis X1

E_Kin Axis X2

a) b)

Figure 3: a) Required kinetic energy; b) Power consumption of control cabinet

201


2 Workpiece clamping via freezing

In order to develop the already mentioned small machine tool, the machine

components, for example workpiece clampings have to be made suitable. Within the

priority program a miniaturized clamping device was developed for fixing a

workpiece via freezing of water [3]. A test workpiece made of 100Cr6 with

dimensions 16 x 16 x 8.5 mm was analysed regarding thermal distortion during the

freezing process. The results show a distortion in z-direction of 5 µm after a freezing

time of 90 s (Figure 4b). Performing higher freezing times showed no modification

regarding the thermal distortion. Using this clamping device the test workpiece could

be fixed on a peltier element within a freezing time of 20 s.

Figure 4: Temperature distribution after freezing time 90 s (a); thermal distortion of

the workpiece (b)

3 Numerical Analysis of the CFRP-frame

In order to evaluate the static stiffness and dynamic behaviour of the machine frame

the FEM-Software ABAQUS was used. The CFRP (Carbon Fibre Reinforced

Plastic) layers were modeled using the “Composite layup”-option. The results

showed a suitable configuration of 15 layers with angles of 0° and 90° [3]. With this

configuration the calculated static stiffness in z-direction was in a range between 95

N/µm up to 120 N/µm depending on the axis position (Figure 5a) by loading the

TCP with experimentally measured feed forces of 1 N in a linear static analysis.

Additionally the magnitude of the simulated dynamic response was analyzed to

0,033 µm/N for Eigenmode 5 at 553,68 Hz and 0,03 µm/N for Eigenmode 7 at

639,34 Hz (Figure 5b). However, when using high speed spindles for machining

there are high rotation speeds, so the operating frequency will be above 2500 Hz

with very low amplitudes (Figure 5b).

202


Figure 5: Static stiffness (a) and magnitude of harmonic response of CFRP-frame (b)

Conclusions

Experimental investigations regarding energy consumption in cooperative mode

show a reduction of kinetic energy of about 50 %. However the power consumption

increases in cooperative mode due to necessary additional drive and electric losses.

Numerical Results show a maximum thermal deformation of 5 µm of the test

workpiece due to the clamping process. The CFRP-frame shows a high stiffness and

low dynamic magnitudes and is suitable for use as machine tool frame also due to its

thermal stability and subsequently higher precision of the machine tool.

Acknowledgement

The authors of this work wish to acknowledge the financial support of the German

Research Foundation (DFG) within the Priority Programme 1476 “Small Machine

Tools for Small Parts”.

References:

[1] Steinhagen, R.: Vorstellung eines Konzepts für den Bau von Sondermaschinen

für die Mikrozerspanung, Workshop zur Fertigung von kleinen Präzisionsteilen für

die Medizintechnik und Analytik, Berlin, 2008.

[2] Wulfsberg, J. P., Grimske, S., Kohrs, P., Kong, N.: Kleine Werkzeuge für kleine

Werkstücke, wt werkstattstechnik online 2010, Ausgabe 11/12, S. 887-891, Internet:

www.werkstattstechnik.de, Springer-Verlag.

[3] Verl, A., Hoffmeister, H.-W., Wurst, K.-H., Heinze, T., Gerdes, A.: Kleine

Werkzeugmaschine für kleine Werkstücke, wt Werkstatttechnik online, Springer-

Verlag, 2012, Ausgabe 11/12, S. 744-749

203


The dynamic design of an ultra-precision machine tool used

for larger KDP crystal machining

Yingchun Liang, Wanqun Chen*, Yazhou Sun, Qiang Zhang, Feihu Zhang

Center for Precision Engineering, Harbin Institute of Technology, Harbin, China

[email protected]

Abstract

This paper presents the design and dynamic optimization method of an ultra-precision

diamond flycutting machine tool for flat surface machining of Potassium Dihydrogen

Phosphate (KDP) crystal in half-meter scale. An accurate multi-degree-of-freedom

dynamic model for this machine tool is built up to describe its static and dynamic

characteristics. The effects of the tool tip response under the cutting force in the

whole cutting path on surface topography and the dynamic structure loop of the

machine tool are analyzed. The weak line of the structure loop is optimized to

improve the dynamic performance of the machine tool. Preliminary machining trials

are carried out, which shows this machine tool can successfully manufacture 430 mm

× 430 mm surfaces on crystalline optics, with 1.3 µm flatness and 2.4 nm Ra

roughness.

1 Introduction

KDP crystal is a kind of crystal material with good nonlinear optical and electro-

optical properties. This crystal is largely applied in the laser fusion system of Inertial

Confinement Fusion (ICF) program as harmonic frequency converters [1]. The KDP

crystal has extremely harsh requirements of the topography in the ICF program. It

requires the flatness less than 3 μm in the whole size of 430×430 mm2, and the

roughness values less than 3 nm Ra [3]. However, this material is so soft, fragile,

gyroscopic, and thermally sensitive that traditional grinding and polishing methods

are not suitable for processing this material, so it's final surface only can be achieved

by cutting. Therefore, an ultra-precision flycutting machine tool urgently is required

to be designed.

204


2 Static and dynamic characteristics analysis

The configuration of the flycutting machine tool is designed as shown in Fig.1. A

bridge supports a vertical-axis aerostatic spindle and flycutter over a horizontal-axis

hydrostatic slide. Mounted to the horizontal slide is a vacuum chuck that fixes the

workpiece by vacuum power. The surface to be machined lays in a horizontal plane.

This configuration can not only improve the rigidity of the machine tool but also

reduce thermal deformation. In order to improve the stiffness of the spindle in the

axial, a large support surface is adopted.

In order to describe its static and dynamic characteristics, an accurate multi-degree-

of-freedom dynamic model for this machine tool is built up as shown in Fig.1. The

tool-workpiece structural loop and the spindle shaft of the machine tool are also given.

Figure 1: The FE model of the machine tool.

2.1 Static analysis

As the cutting proceeding, the cutter moves from point A to point B as shown in Fig.1.

This cutting process is simulated by the Finite Element (FE) method. The results

show that the tool tip has different displacement in z direction, with the cutting force

is 1 N, 10 N and 10 N in x, y, and z direction, respectively. The displacement of the

tool tip in z direction is shown in Fig.2, the maximum value up to 1 μm at point C,

which will result in a convex surface. This phenomenon is because that the spatial

position and direction of the cutting force change constantly along the whole cutting

path. It indicates that to obtain a flat surface, the spindle axis should be located

slightly forward to the slide, rather than vertical completely, when installing the

spindle.

205


Figure 2: The displacement of the tool tip during the cutting proceeding.

2.2 Dynamics analysis

In order to improve the dynamic performance of the machine tool, analysis of the

gantry structure machine tool in detail is given as follows. The contribution of the

crucial machine components to the dynamic performance of the tool-workpiece

structural loop is given in Fig.3. The response points are laid on the tool tip and

workpiece, respectively. It shows that the first order mode of vibration of the tool-

workpiece structural loop has the same mode shape with the machine structure, but

the value is less than the machine structure’s. The first order frequency of the spindle

and slide occurs at 324 Hz and 425 Hz, respectively. It demonstrates that the spindle

and the slide have a good dynamic performance; The weak link within the tool-

workpiece structural loop is the machine structure, which has the most significant

effect on the dynamic performance of the machine tool. That's because the machine

structure provides the support and accommodation for the spindle component, the

additional weight of the spindle makes the dynamic performance of the machine

structure decrease sharply, the first order frequency decreases from 164 Hz to 105 Hz.

Therefore, in order to improve the dynamic performance of the machine tool, the

machine structure is optimized as follows.

The workspace provided by the machine structure is designed as 770×518 mm, thus it

can meet the requirement of both machining the 430×430×10 mm workpiece and

giving enough space for the spindle, slide and the vacuum chuck. The objective

function is simply defined as the maximum value of the first order frequency of the

machine structure. The limitation is the workspace provided by the machine structure,

and the design variables are the structural parameters of the machine structure. After

optimization, the first order frequency of the machine structure increases from 222 Hz

206


to 325 Hz. The dynamic performance of the tool-workpiece structural loop rises from

116 Hz to 145 Hz.

Figure 3: The influence of the components of the machine tool on the tool-workpiece

structural loop. ① The first mode of the machine structure; ② The second mode of the

machine structure; ③ The first mode of the spindle; ④ The first mode of the slide; ⑤

The first mode of the tool-workpiece structural loop; ⑥ The second mode of the tool-

workpiece structural loop.

3 Preliminary machining trials on the machine tool

The machine tool has been utilized to machine the KDP crystal with size of 430×430

mm. The preliminary machining results had a roughness of 2.4 nm Ra and a flatness

of 1.3 µm as shown in Fig.4.

Figure 4: The test results.

References:

[1] Painsner JA, Boyes JD and Kuopen SA. National ignition facility. Laser Focus

World 1994; 30: 75.

[2] Lahaye P, Chomont C and Dumont P. Using a design of experiment method to

improve KDP crystal machining process. Proc SPIE 1998; 3492: 814–820.

207


Investigation of micro-optic polishing characteristics by

vibration-assisted polishing

J. GUO1*, Y. YAMAGATA 1, H. SUZUKI1, 2, S. MORITA 1 and T. HIGUCHI3 1The Institute of Physical and Chemical Research (RIKEN), Wako, Saitama, Japan 2Department of Mechanical Engineering, Chubu University, Kasugai, Aichi, Japan 3Department of Precision Engineering, The University of Tokyo, Tokyo, Japan

*[email protected]

Abstract

The micro-optic polishing characteristics are investigated. The material removal rate

and surface roughness under different vibrating motions are compared. The

relationship between polishing pressure and material removal rate is revealed. It is

found that the result does not follow the Preston’s equation completely because the

material removal rate decreases when the polishing pressure exceeds a certain value.

A model of material removal mechanism in micro-optic polishing is proposed and

illustrated.

1 Introduction

Recently, the vibration-assisted polishing method has been proposed to finish the

micro-optic mould and some good results have been reported [1-2]. To well control

the polishing performance and investigate the material removal mechanism, in this

paper, some fundamental experiments are conducted to investigate the micro-optic

polishing characteristics. Firstly, the experimental setup and polishing conditions are

illustrated. Then the material removal rate and surface roughness under different

vibrating motions are compared to find the suitable polishing condition. After that,

the relationship between polishing pressure and material removal rate is revealed.

Finally, a model of material removal mechanism in micro-optic polishing is proposed.

1 Experimental setup and polishing conditions

The experimental setup for micro-optic polishing characteristics investigation is

shown in Fig. 1. It was presented in the previous research which consists of a

magnetostrictive vibrating polisher, a real-time polishing force control system and a

208



5-axis NC machine [2]. Major polishing conditions are summarized in Table 1. The

workpiece made of tungsten carbide is lapped prior to polishing. Tool dwell control

method by Zigzag scanning is adapted to polishing experiments as shown in Fig. 2.

Figure 1: Experimental setup for polishing Figure 2: Tool dwell control method

Characteristics investigation by Zigzag scanning

Table 1 Main polishing conditions

Workpiece material Binderless tungsten carbide

Polisher head

Radius

Hardness

Polyurethane

1 mm

IRHD 90

Abrasive

Grain size

Density

Diamond slurry

0.1 μm

1 wt%

Polishing forces 5.0 - 50.0 mN (Increment: 5.0 mN )

Polishing scope 400 × 400 μm2

Scanning speed 3.5 mm/min

Pitch size 20 μm

2 Relationships between polishing parameters

2.1 Material removal rate and surface roughness

Some experiments are conducted to compare the material removal rate and surface

roughness under different vibrating motions. The magnetostrictive vibrating polisher

209


scanned on the work-piece with the lateral, elliptical (phase difference of 45 deg)

and circular vibrating motion, respectively [2].

The polishing force is set to 5 mN and the grain size of diamond slurry is opted to

0.1 μm with polisher radius of 1 mm. The data is measured by NewView 6000

(Zygo Corporation) and the results are summarized in Table 2. It is proved that

circular vibrating motion has the highest polishing efficiency with the removal depth

up to 50 nm/min.. The surface roughness is reduced over 50% by the 2D vibrating

motion such as elliptical and circular vibration than that of the 1D or lateral

vibrating motion..

Table 2 Removal depth and surface roughness under different vibrating motions

Vibrating motion Removal rate*1 Surface roughness*2

Lateral 30 nm/min. 8 nm Rz (1 nm Ra)

Elliptical 40 nm/min. 3.4 nm Rz (0.35 nm Ra)

Circular 50 nm/min. 3.1 nm Rz (0.4 nm Ra)

*1 Average in 10 times *2 Measurement area size: 70 x 50 μm2

2.2 Polishing pressure and material removal rate

Further investigations are conducted to check the relationship between polishing

pressure and material removal rate. The circular vibrating motion and the grain size

of 0.1 μm of diamond slurry are adopted.

The results are shown in Fig. 3. It is found that the material removal rate shows

agreement with Preston's equation when the polishing pressure is under 345 kPa.

But when the polishing pressure exceeds 345 kPa the removal rate decreases

gradually with the increasing of polishing pressure.

In order to explain this phenomenon, a model of material removal mechanism for

micro-optic mould polishing is proposed for the first time in this research as shown

in Fig. 4. Although the fundamental material removal mechanism is poorly

understood and a holistic knowledge still does not exist since the physical scale of

material removal processes in polishing is difficult (practically impossible) to be

observed directly [5], there is a general view that two kinds of abrasive motion

which are two-body abrasion and three-body abrasion effect on the work-piece

during loose abrasive polishing process. Two-body abrasion happens when abrasives

210

http://www.zygo.com/


become embedded and slide over the surface, while three-body abrasion is generated

when abrasives become freely rolling abrasives [3]. In this experiment, the urethane

polisher is soft and the polishing spot which is the area of polishing removal

function is very small (under 0.2 mm2). So when the polishing pressure is over 345

kPa the number of two-body abrasives between the polisher and work-piece

decreases with the polishing pressure. The two-body abrasives are dropped out due

to the high pressure and the number reduces, although maybe some of them

transform to the three-body abrasives, the total material removal rate decreases.

Figure 3: Relationship between polishing Figure 4: Material removal mechanism in

Pressure and material removal rate micro-optic mould polishing

3 Conclusions

According to the above-mentioned experiment results, it can be concluded that in

case of micro-optic mould polishing, there is a certain value of polishing pressure

which the material removal rate researches to a maximum. In other words, the

material removal rate does not always increase in linearity with the polishing

pressure, and when the polishing pressure exceeds a certain value, it decreases. So it

will be a great complement to the application of Preston’s equation in micro-optic

polishing.

References:

[1] H. Suzuki, et al., 2010, Annals of the CIRP, Vol. 59/1, pp. 347-350.

[2] J. Guo, et al., 2012, Annals of CIRP, Vol. 61/1, pp. 371-374.

[3] E. Brinksmeier, et al., 2006, Precision Engineering Vol.30/3, pp. 325–336.

[4] C.J. Evans, 2003, Annals of the CIRP, 52/2, 611-633.

211


Parameter Determination for an Electromechanical Model

of a Displacement-Amplified Piezoelectric Actuator

J.H. Liu1, W.O’Connor1, E.Ahearne1 and G.Byrne1 1 School of Mechanical and Materials Engineering, University College Dublin,

Belfield, Dublin 4, Ireland

[email protected]

Abstract

The “long range displacement” piezoelectric actuator (PEA) investigated in this paper

comprises pre-stressed piezoceramic lead zirconate titanate (PZT) stacks in a flexure-

constrained multi-component frame. This paper proposes a PZT electromechanical

model which relates the stacks' electrical and mechanical domains. This paper

introduces an identification approach to the determination of the model parameters

without disassembling the embedded piezoceramic stacks. The electromechanical

couplings of the PZT stacks, which describe the energy transfer between the electrical

and mechanical domains, were experimentally identified.

1 Introduction

Our research group is pioneering the use of long range PEAs for closed loop control

of the applied force in a number of manufacturing processes including chemical

mechanical planarisation (CMP), with the potential to provide a major improvement

in the control of the local interfacial pressure between the silicon wafer and polishing

pad [1]–[2]. The PEA is a commercial product, “Flextensional Piezoelectric

Actuator™” (Dynamic Structures and Materials, LLC, Franklin, USA) which is

composed of PZT stacks and a flexure-hinged amplification mechanism (FAM). The

PZT layers are electrically connected in parallel generating strain when charged

which is magnified by the flexible mechanism so as to realise a relatively large

output displacement, as showin in Fig. 1 (a).

2 Methodology

In the PEA model, the whole actuator was divided into the PZT stacks, the

mechanical parallel pre-stress springs, and the external FAM. In this approach the

212



electromechanical model of the pre-stressed multi-layer PZT stacks is supplemented

with a kinematic model of the FAM.

2.1 Electromechanical model of PZT stacks

The electromechanical model of the PZT stacks is shown in Fig. 1 (b). The input to

this model is voltage, and the output is the PZT displacement. The total input voltage

(vin) was divided into the voltage that induces hysteresis (vh) and the voltage linearly

proportional to the piezoelectric force (vp). H represents the hysteresis operator, x is

the stack displacement, fp is the transduced force from the electrical domain. The

electrical and mechanical domains are related by the electromechanical coupling

factors. T is the electromechanical coupling between piezoelectrical charge and

displacement, and N is the factor between voltage and displacement. In the PZT

model a linear relationship is assumed between the mechanical and electrical

domains:

qp=T×x Eq. 1

vp=N×x Eq. 2

The total current is the sum of the current through the capacitor, resistor and the

current introduced by the piezoelectric effect. So the charge can be defined by:

q=c×vp+ qp + qr* Eq. 3

* dynamic part only.

The mechanical part was modelled as a linear, lumped mass-spring-damper system.

The equivalent capacitance c, equivalent resistance r, the equivalent mass mp,

damping ratio bp, the stiffness of PZT stacks kp, and the stiffness of preload springs ks

are parameters that need to be identified.

Output Connection

with Mounting Holes

Spring Preload

Strap

Stainless Steel

Flexure Frame

Base Mounting

Connection

Multi-layer

PZT Stack

Direction of

Motion

(a)

vin

mp

ks bp

x

kp

H

c r

vh

vp

fp

q

cq rq pqT

N(b)

Fig. 1: (a) DSM PEA; (b) Electromechanical model of the PZT stacks

213


2.2 Parameters identification

During identification two PEAs were fixed in series in a closed structural loop in a

Hounsfield testing machine, as shown in Fig. 2. The preload force exerted on the

PEAs is adjustable. In the experiments, the actuator #2 was investigated in short and

open circuit conditions. Under both conditions, tensile and compressive forces were

applied by actuator #1. The induced displacements in horizontal directions were

measured while the vertical force was measured by an inline force sensor.

Load CellFixing Screw

PEA #1

ConnectingScrew

Connecting Screw Dis.

Sensors

PEA #2

HounsfieldFrame

PEAAmplifier

Current Flow

Force Sensor

Hemisphere

Pivot

(lubricated)

Fig. 2: Experimental setup: actuator #2 in short circuit condition

The principle of this experiment is that, when the actuator was tested in short circuit,

the coupling between the charge and the PZT stack displacement can be identified,

and when the actuator was tested in open circuit, the coupling between the voltage

and PZT stack displacement ( or force) can be determined.

3 Results and discussion

(a) 0 0.2 0.4 0.6 0.8 1

0

5

10

15

20

Time [Sec]

Voltage [

V]

(b) 0 0.5 1 1.5

-8

-6

-4

-2

0

x 10-5

Open Circuit Voltage [V]

Short

Sircuit C

harg

e [

C]

Fig. 3: (a) Signal to PEA #1; (b) Short circuit charge versus open circuit voltage

The driving signal to PEA #1 is shown in Fig. 3 (a). The relationship between the

short circuit charge and the open circuit voltage is shown in Fig. 3 (b) indicating that

hysteresis is not relevant for this situation. The open circuit voltage and short circuit

214


current is shown in Fig. 4 (a) and (b). The slope of Fig. 5 (a) represents the coupling

ratio between PZT displacement and the linear voltage. The slope of Fig. 5 (b)

represents the coupling ratio between short circuit charge and the PZT displacement.

The identified parameters are listed in Table1.

(a)0 0.2 0.4 0.6 0.8 1

0

0.5

1

1.5

2

Time [Sec]

Open C

ircuit V

oltage [

V]

(b)0 0.2 0.4 0.6 0.8 1

-2

-1

0

1

2x 10

-3

Time [Sec]

Short

Circuit C

urr

ent

[A]

Fig. 4: (a) Open circuit voltage of PEA #2; (b) Short circuit current of PEA #2

(a) 0 0.5 1 1.5

-10

-5

0

x 10-7

Open Circuit Voltage [V]

PZ

T D

ispla

cem

ent

[m]

(b) -1.4 -1.2 -1 -0.8 -0.6 -0.4 -0.2 0

x 10-6

-10

-8

-6

-4

-2

0x 10

-5

PZT Displacement [m]

Short

Sircuit C

harg

e [

C]

Fig. 5: (a) PZT displacement vs. open circuit voltage; (b) Short circuit charge vs.

displacement

Table1: Identified electromechanical coupling coefficients

Parameters N Units T Units

Value -1.76×106 [V/m] 67.62 [C/m]

In summary, in this paper the values of the coupling ratios between electrical and

mechanical domains in the PZT model were experimentally determined and these

values will be used in future modelling work.

References:

[1] J. Liu, E. Ahearne and G. Byrne, “Characterisation of the Transfer Function of

an Advanced Process Control System for Chemical Mechanical Polishing

(CMP),” in Proc. 2011 11th International Conference of the European Society

for Precision Engineering & Nanotechnology, Como, Vol.1, 2011, pp.311-314.

[2] J. Liu, E. Ahearne and G. Byrne, “Characterisation of the External Loading

Conditions of an Advanced Process Control System Integrated with

Piezoelectric Actuator (PEA) in Chemical Mechanical Polishing (CMP),” in

Proc. 2011 28th International Manufacturing Conf., 2011, Dublin , pp.1-8.

215


Ultraprecise positioning mechanism with 3-DOF over a one-

millimeter stroke using monolithic flexure guide and

electromagnetic actuator

S. Fukada1, T. Matsuda, Y. Aoyama, T. Kirihara 1Shinshu University, Japan

[email protected]

Abstract

A new mechanism is introduced by integrating monolithic flexure mechanisms, and

the performance of the mechanism is discussed. Ultra-precise circular motion with 1

mm diameter is achieved with 8.2 nm (P-V value) deviation from circularity.

1 Introduction

Current precise positioning mechanisms can be divided into two categories based on

their field of application: The first category is positioning mechanisms with long

strokes from millimeters to meters; the second category is fine positioning

mechanisms with strokes measured in micrometers. To meet a medium range

between these two categories, the authors attempted to create a positioning

mechanism with nanometric resolution over a 1-mm stroke using a flexure guide and

an electromagnetic actuator: They had previously reported a planer positioning

mechanism with three degrees of freedom (3-DOF), in which both ultraprecision and

ultrafine point-to-point (PTP) positioning with resolution of 2 nm for X–Y and 0.01

asec for (yawing) was achieved over a 1-mm stroke [1, 2]. However, because the

mechanism was constructed using 32 pieces of leaf springs, there was some

interference between X-Y- axes of motion that deteriorated the positioning

performance in continuous-path (CP) motion. In this report, a new mechanism is

fabricated by integrating monolithic flexure mechanisms, and the performance of the

mechanism under multi axis control for CP positioning is discussed.

2 Mechanism

Figure 1 shows the developed mechanism. It consists of two pieces of a monolithic

flexure device, which forms a positioned stage of cube configuration with 60 mm

sides, and flexure thin beams of double compound rectilinear springs supporting the

216


stage [3]. These parts were shaped from a stainless steel plate using wire electric

discharge machining. Three pairs of voice coil motors are set around the stage, and

these motors independently generate driving forces or yaw moment in the X, Y and

directions. An optical square is placed on the table to measure stage motion in the X-

Y- directions using a laser interferometer with resolution of 0.6 nm. The static

compliance of the flexure guide was 112 m/N in X, 81 m/N in Y and 0.011 rad/Nm

in : The stage can move approximately 0.4 mm/A. To compensate for any damping

effect, the mechanism is sunk in silicone oil of 1000 cSt.

First, quasi-static and dynamic characteristics of the mechanism as a multiple-input

multiple-output (MIMO) system were determined. Figure 2(a) shows quasi-static

response to ramp-input of the mechanism: The light lines show the natural property of

the plane mechanism. The mechanism has superior properties with a linear relation

between the input and the response of corresponding axis as shown in diagonal

elements of the figure. The interference between the axes as shown in off-diagonal

elements is greatly decreased compared with the previous mechanism. Figure 2(b)

shows frequency response of the mechanism.

3 Control system

Figure 3 shows the multi-axis control system. To counter the slight residual

interference among axes shown in Fig. 2 by light lines, a decoupling compensator is

(a) Schematics (b) Arrangement of flexures, actuators and sensors

Figure 1: Positioning mechanism

217


applied. The properties existing in off-diagonal elements of the transfer matrix G of

the mechanism were identified from the quasi-static characteristics measured in Fig.

2(a): The inverse matrix G-1 was calculated in order to eliminate the interference. The

matrix Cr in the figure represents ideal reference gains of the controlled object. The

quasi-static and dynamic responses to operating variables V are shown in Fig. 2 by

dark lines: The interference of the mechanism has been greatly reduced in both quasi-

static and dynamic conditions. A two-degrees-of-freedom control system with

feedback PID controller and feedforward compensator Cr-1 is realized by using a

digital signal processor with sampling rate of 3 kHz.

4 Continuous path control result

Next, performance of CP circular motion was determined with simultaneous control

of all axes of X-Y- . The reference Xr is a sine wave function, and Yr is a cosine wave

(a) Quasi-static characteristics (b) Dynamic characteristics

Figure 2: Input-output response of the MIMO system

Figure 3: Multi axis control system

218


function; r is controlled to keep the yawing

motion at zero. Figure 4 shows the controlled

deviation of each axis in circular motion with 1

mm diameter [4]. The controlled deviation of X

and Y axes shows accurate performance, with

tracking error for each axis less than 2 nm

(RMS). Here, the decoupling compensator

showed its effect in the wide range of frequency.

Figure 5 shows deviation from roundness of the obtained circular motion: Ultra-

precise circular motion with 1 mm diameter is achieved with only 8.2 nm (P-V value)

deviation from circularity.

5 Conclusion

A new mechanism is fabricated by integrating monolithic flexure mechanisms, and

the performance of the mechanism under multi axis control for CP positioning is

discussed. The superior performance of the developed system is verified, and the

potential of the developed mechanism for CP motion is clearly demonstrated.

References:

[1] S. Fukada, et al.: Proc. euspen 9th Int. Conf., (2009) p. 341-344.

[2] S. Fukada, et al: Int. J. Automation Technology, Vol. 5, No. 6 (2011) p. 809-822.

[3] S.T.Smith,D.G.Chetwynd:“Foundation of Ultraprecision Mechanism

Design, ”Gordon and Breach Science Publishers, (1994).

[4] L. Jabben, J van Eijk: Mikroniek, Vol. 51, Issue 3 (2011) p. 16-21.

(a) Tracking deviation of each axis (b) CAS of tracking deviation

Figure 4: Tracking performance of each axis

Figure 5: Deviation from

roundness

219


Design and construction of a novel assisted tool-holder

L. Javarez Jr 1, J.G. Duduch1, R.G. Jasinevicius1, A.M. Gonçalves1 1University of São Paulo, Brazil

[email protected]

Abstract

This paper presents a novel design of an Assisted Tool-holder (ATh) for ultra-

precision single point diamond machining. The combination of a piezoelectric

actuator to produce displacements, a non-contact capacitive sensor to accurately

measure these displacements and a PID control system that maintains the accuracy of

the displacement of the tool actuator ensures the quality of machining. Also, the

design complies with the principle of symmetry and makes Abbe offset as small as

possible. Backlash is avoided with the use of flexures at the operating range of 0-30

μm. Finite Element Method (FEM) analyses including strain stress due to forces

acting on the system is employed. Fatigue analysis was performed to predict the

lifetime of the ATh, and, finally, nodal analysis was performed to predict the natural

frequencies of the system. The quality of the model was accomplished by optimizing

the mesh according to the Skewness Criterion. Analyses took into account the

reaction force of the flexure on the actuator due to the maximum displacement of 30

μm. The results of preliminary simulations showed that the ATh meets all the

requirements of resolution, frequency response and compactness.

1 Introduction

Most experiments concerning micro and nano-positioning rarely meet a real

application in precision engineering [1]. On the contrary, isolated tests have been

done without control of the cutting force and designs seldom follow any evaluative

mechanical principle (fatigue, effects of stress-strain response).

An example of a tool holder designed to achieve a maximum displacement of 7.5 μm

at 100 Hz is described in [2]. In order to increase the displacement of this actuator,

an amplifier was incorporated into the piezoelectric mechanism and the displacement

was increased to 432 microns, compromising however the fatigue life. An interesting

study on stiffness in three directions (x, y, z) of a fast tool servo is presented in [3].

220



This work aims to design an Assisted Tool-holder using Ansys® FEM analysis

simulation tools ensuring the model’s effectiveness by the Skewness criterion and

checking the reaction forces caused by a maximum displacement of 30μm.

2 ATh design

The prototype of the Assisted Too-holder, ATh, was designed and constructed as

shown in Figure 1. The device is composed of 6 main parts. A flexural bearing of the

monolithic type which generates pre-loading and restoring force, a sensor mount, an

actuator mount, a tool-holder, a piezoelectric actuator and a capacitive sensor. The

design adheres to the principles of alignment and symmetry and contains no moving

parts. Also, the piezoelectric actuator is firmly attached to the tool-holder, eliminating

noises and imperfect contacts.

Figure 1: Schematic view of the ATh parts: (1) actuator mount, (2) flexural bearings,

(3) tool holder (4) sensor mount, (5) capacitive sensor, and (6) piezoelectric actuator.

3 Generation and optimization of the FEM mesh and results

The Skewness criterion was used to evaluate and optimize the effectiveness of the

mesh. Its main function is to quantify how close an element in the mesh is to the

ideal, so it quantifies the distortion of the actual pattern of the element. This method

is defined as:

Skewness = (optimal size of element) – (element size)

(optimal size of element)

221


The average skewness value of the elements was 0.527. This criterion ranges from 0

to 1 (excellent to bad, respectively). Therefore this mesh can be considered good.

Special attention was given to contacts considered critical (flexures/tool-holder) to

refine the mesh, as observed in Figure 2. Entire mesh study is needed to ensure the

fidelity of the model relative to the actual model.

Reaction force value corresponds to a displacement of 30 μm and 0.5 mm spring

thickness in aluminum alloy 7075. Figure 3 shows the average von misses stresses

caused by this displacement.

Figure 2: Mesh refinement.

Simulation of nodal frequencies showed the first natural frequency to be above 1

kHz for 1.5 spring thickness. Fatigue analysis predicted a lifetime of 1011 and 1010

for spring thickness of 1.5 and 0.5, respectively. A simple, symmetric design proved

itself to attend the requirements of fatigue life and frequency response.

Figure 3: Von Misses stress

222


Table1: Results of life and natural frequencies

Material Thickness of

spring (mm)

Life

(cycles)

Natural frequency

1st 2nd 3th

Aluminium

7075

1,50 1011 1008,8 1535,4 1742,8

0,50 1010 754,7 932,1 1297,3

4 Conclusions

An assisted piezoelectric tool holder was designed and constructed. An optimized

mesh (skewness criterion) Finite Element model was used to predict displacement,

lifetime, stress and natural frequency. Simulations showed that the FEM model and

the prototype attend all requirements of resolution and frequency response.

References:

[1] Li, S. Z., Yu, J. J., Pei, X., Su, H. J., Hopkins, J. B., And Culpepper, M. L.: Type

synthesis principle and practice of flexure systems in the framework of screw

theory: Proceeding of the ASME 2010 international design engineering technical

conferences & computers and information in engineering conference

IDETC/CIE 2010, Montreal, Quebec, Canada, 15–18 August 2010, DETC2010-

28963, 2010.

[2] Kim, H. S., Kim, E. J., And Song, B. S. Diamond turning of large off-axis

aspheric mirrors using a fast tool servo with on-machine measurement, J. Mater.

Process. Tech., 146, 349–355, 2004.

[3] Gan, S. W., Lim, H. S., Rahman, M., And Watt, F.: A fine tool servo for global

position error compensation for a miniature ultra-precision lathe, Int. J. Mach.

Tool. Manu., 47, 1302–1310, 2007.

[4] Tian, Y., Shirinzadeh, B., And Zhang, D.: A flexure-based mechanism and

control methodology for ultra-precision turning operation, Precis. Eng., 33, 160–

166, 2009.

223


Development of a Vertical-spindle Rotary Surface Grinding

Machine for Large Scale Silicon-wafers – Machine

Specifications and Performance of Rotary Work Table

A.Yui1, A.Honda1, S.Okuyama1, T.Kitajima1, G.Okahata1, H.Saito2 and A.H.Slocum3

1National Defense Academy, Japan2Okamoto Machine Tool, Japan3Massachusette Institute of Technology, USA

[email protected]

Abstract

The development of a next generation surface-grinding machine for 450mm

diameter silicon-wafers is required from the semiconductor industry. Loop stiffness

of the grinding machine has to be high enough to sustain high grinding force

because of the large contact area between the grinding wheel and the wafer. To

increase the loop stiffness of the machine, each machine component should have

high stiffness; the number of the components should be as small as possible and,

thus, the machine construction should be simple. The authors developed a new

vertical-spindle surface grinding machine equipped with a rotary work table

sustained by water hydrostatic bearings, a wheel spindle equipped with a wheel

infeed system and a kinematic cupping system that

firmly fixes the wheel spindle head against the

base of the work table1). This paper describes the

specifications of the developed grinding machine

and investigates the results of static stiffness and

rotational accuracy of the work table. Measured

static stiffness of the work table was 2.5kN/m

under water flow rate of Q=10mL/min and

rotational accuracy was 0.25m under 120rpm.

1 Machine specification and construction

of rotary table

Figure 1 shows a photograph of the developedFig.1 Photograph of developed

grinding machine for 450mmdiameter

Rotary work table

Bed

Wheel spindle

Positioning sensor

Kinematic coupling

224


grinding machine and Table 1 shows the specifications of the machine. Constant

water flow is supplied to the thrust bearings of the work table using micro-gear

pumps, and constant water pressure is supplied to both the radial bearings of the

work table and the wheel spindle. The wheel spindle is axially sustained by a linear

motor. The hydrostatic bearings are earth-friendly, because the working fluid of the

bearings is pure water. A porous chuck is installed on the table to vacuum hold a

silicon-wafer.

Figure 2 shows a schematic

of the rotary work table which is

axially sustained by a single recess

type constant flow hydrostatic

water bearing2)3). Strong

neodymium magnets are installed

under the rotary table to preload

the table in an axial direction

(6kN) and thus reinforce the

bearing stiffness. The bearing pad

is optimally designed to realize the

necessary sustaining force and the

static stiffness of the rotary work

Table 1 Specifications of developed rotary surface grinding machine

Mai

nB

ody

Machine size 2000×2000×2400 (mm)Wafer diameter 450mmBearing type Hydrostatic bearingWorking fluid Pure water

Wo

rkta

ble

Table diameter 500mm

Table mass 300kgTable rotational speed 0-500rpm

Bearing typeThrust Constant flow hydrostatic, Q=10-50mL/min

Radial Constant pressure hydrostatic, P=1.2MPa

Wafer clampingmethod

Vacuum porous chuck

Wh

eel

spin

dle

Rotational speed 0-2500rpmFeed stroke 1.5mmFeed speed 0.010-0.999mm/minMinimum increment 10nm

Bearing typeThrust Linear motor

Radial Constant pressure hydrostatic, P=1.2MPa

Fig.2 Schematic of rotary work table equippedwith a water hydrostatic bearing

Thrust bearing

Neodymium magnets

Direct drive motor

Radial bearing

Rotary work table

225


table. Flow rate of each bearing Q can be changed by adjusting the rotational speed

of the micro gear pump. The rotary table is radially sustained by constant water

pressure hydrostatic bearings (the pressure P=1.2MPa) and the rotational speed is

controlled using a direct-drive servo motor.

2 Experimental results of work table performance

Figure 3 shows the effect of flow rate Q of each bearing on bearing gap h. The

measured data is plotted in Figure 3 with the curve being the derived value from

equation (1).

30

3 'CC hhhQ (1)

where C is a constant given from table mass, bearing preload, viscosity and effective

bearing pad size, h0 is gap margin and h’ is effective bearing gap. In this system, the

table does not float (h’=0) until

Q=0.8mL/min. This is due to the

assumption that some water leakage

occurs from the bearing surface.

Therefore, h0 must be considered in

calculating the real bearing gap h.

Fig. 4 shows the effect of

Q on static stiffness K of the work

table. Weights of 18.5kg mass are

placed on the table one by one and

the vertical displacement of the

table is measured using three

electric micrometres. Lower Q

results in higher K, and the

measured K under Q=10mL/min

(h’=6m) was 2.5kN/m. The

bearing stiffness is high enough to

compose the high precision

grinding machine table.

In the case of rotary Fig.4 Effect of flow rate Q on static stiffness K

Flow rate Q mL/min

Sta

tic

stif

fnes

sK

kN

/m

10 20 30 40 50 6001

1.5

2

2.5

3

10 20 30 40 50 60

5

0

5

10

15

20

Fig.3 Effect of flow rate Q on bearing gap h

20

10 20 30 400

-5

5

10

15

Q mL/min

Bea

ring

gap

h,h

’μ

m

50 60

h’

h0

h

226


grinding of silicon-wafers, normal

grinding force is not applied on the

center of the rotary table, but it is

applied eccentrically. Fig.5 shows the

effect of eccentric force on table

inclination angle . When a moment of

181N·m is applied on the table under

Q=50mL/min, measured was 3.94".

This shows that when the load of 100N

is applied to the outskirts of the table

(table radios=250mm), the loading point

will sink 0.61m, which is stiff enough

to grind silicon wafers.

Figure 6 shows a radial

motion-deviation of the work table under

120rpm and Q=10mL/min. The radial

deviation is measured using a master ball

(0.055m roundness in a measuring

plane) and electric micrometers, which

are set perpendicularly in the measuring

plane. Measured rotating accuracy was 0.25m.

3 Conclusions

A next generation precision grinding machine for 450mm diameter silicon-wafers is

developed and performance of the work table is investigated. Measured static

stiffness of the table was 2.5kN/m under Q=10mL/min, and rotational accuracy

was 0.25m under 120rpm.

References:

[1] G.Rothenhöfer, A.H.Slocum, M.Paone, X. Lu, A. Yui: Proc. of 11th euspenInternational Conference, (2011.5) pp.260-266.

[2] G.Rothenhöfer, A.H.Slocum, A. Yui: Proc. of 10th euspen InternationalConference, (2008.5) pp.509-513.

[3] A.Yui, S.Okuyama, T.Kitajima, E.Fujita, A.H.Slocum and G. Rothenhöfer: Proc.of the 9th euspen International Conference, (2009.6) pp.248-251.

Fig.6 Radial motion deviation of rotarytable (120rpm, Q=10mL/min)

m

200 400 800 1000 12000

5

10

0

15

20

25

Fig.5 Effect of eccentric force on tableinclined angle θ (Q=50mL/min)

Eccentric load N·m

An

gle

of

incl

inat

ionθ

"

"

600

227


Band-limited Cutting Force Control in Ultra-precision

Turning

K. Enomoto1, Y. Kakinuma1

1Department of System Design Engineering, Keio University, Japan

[email protected]

Abstract

Ultra-precision cutting has recently attracted attention to produce optical parts such

as lenses without a grinding and a polishing process. It is well known that unstable

fluctuation in cutting force has a significant influence on a machined surface

roughness and shape quality. Therefore, it is considered essential to control the

cutting force in the ultra-precision machining. The sensor-less cutting force control

based on the disturbance observer is a practical approach because it does not require

any additional sensors. A cutting force is estimated by using the servo information.

When the estimated cutting force is fed back to the controller, the sensor-less

cutting force control is realized. In this study, the sensor-less cutting force control is

applied at a certain range of frequency, and a position control is simultaneously

employed. The effect of the proposed control method is evaluated by face turning

tests.

1 Introduction

At present, the manufacture of molds and lenses has needed grinding and polishing

after cutting process in order to improve the quality of the machined surface. On the

other hand, for reducing the production cost and raising the energy efficiency, the

ultra-precision cutting has recently attracted attention to produce these parts without

grinding and polishing processes1,2). In this study, we have focused on the behaviour

of cutting force in ultra-precision machining. Altintas3) proposed to control the

cutting force with dynamometer in the millimetre-scale machining. He had shown

that the cutting force control improved the shape accuracy. As well as the

conventional machining, it is considered to be essential to control the cutting force

in the ultra-precision machining. As a practical technique to control the force, the

sensor-less cutting force control is available 4). However, in case that the force

228


control is used for the cutting process, the machining time could not be estimated. In

this study, even as the position control is basically employed, a cutting force at a

certain range of frequency is attempted to be controlled. The effect of the proposed

cutting force control is experimentally evaluated by using the prototype ultra-

precision turning machine.

2 Sensor-less cutting force control system design at a certain range of

frequency

Figure 1 shows applied forces in feed direction

to the X stage during face turning. Based on

the motion equation of the X stage, a cutting

force in feed direction ௨௧ܨ is estimated from

the current reference ܫ

, the position

response ௦andݔ the friction force at the guide

ܨ , as shown in Eq. 1.

௨௧ܨ = ܫ௧ܭ

− ሷݔܯ௦− ܨ (1)

:௧ܭ Nominal thrust force coefficient, :ܯ nominal mass of carriage.

The friction force needs to be identified by idling tests in advance. When the

estimated cutting force is fed back to the controller, the sensor-less cutting force

control is realized. To retrieve the cutting force information at a certain range of

frequency, a band-pass-filter (BPF) is set. To avoid the interaction between a

position control and a force control, the position controller is designed as its

wideband width is lower than that of the BPF. Then, the position controller and the

force controller are integrated at the acceleration dimension, as shown in Fig. 2. aref

represents the acceleration reference. To verify the validity of the designed control

Figure 2 Band-limited force control system

Table 1 Control parameters

Figure 1 Forces applied to X stage

229


Figure 3 Simulation results of the control method

system, a simulation is

performed under the

condition listed in Table 1.

Figure 3 shows the results of

the velocity response and the

estimated force after passing

through the BPF. The

average velocity response

almost corresponds to the

command value and the force response is obviously fluctuated at 10 Hz according to

the force command. It is confirmed that a position and a cutting force can be

simultaneously controlled by separating both frequency ranges without overlaps.


Figure 4 shows the prototype ultra-precision

turning machine which consists of the work

spindle supported by aerostatic bearing and

the linear motor driving carriage. Because

the work spindle employ the non-contact

mechanism, its friction force is almost zero,

which enhance the accuracy of the sensor-

less cutting torque control. In terms of the

carriage, the friction force at the LM guides

is identifed according to each position by the idling test. The carriage is driven by

the linear motor and the optical linear encoder with 10nm resolution is attached at

the side. The position-force-integrated contorol system is installed to the turining

machine. The same control parameters are set as Table 1.

4 Experimental result

To investigate the performance of the proposed control method when a variable load

is applied in feed direction to the tool set on the XZ-stage, the behaviours of the

position and the band-limited force are experimentally evaluated by giving the

velocity and force commands shown in Table 1. Figure 5 shows the relation

Figure 4 Prototype of ultra-precision turning machine

230


between the position and the

force in the x-direction. A

quasi-static force is changed in

accordance with the applied

load because the position of the

stage is controlled within the

bandwidth of 5Hz. On the other

hand, it is clear that the force

ranging from 5Hz to 15Hz is

effectively controlled. From the result, there is possibility to improve cutting

process in terms of the cutting force and the machined surface quality by applying

the proposed hybrid control method. We are planning to show some results of

turning tests on site.

5 Conclusion

A band-limited cutting force control method is proposed for the ultra-precision

turning process. The validity of the proposed method is confirmed by the simulation

and the experiment. In future work, optimum parameters to enhance the turning

process will be investigated.

Acknowledgement:

This study was supported by the Industrial Technology Research Grant Program in

2009 from the New Energy and Industrial Technology Development Organization

(NEDO) of Japan and JSPS KAKENHI Grant Numbers 24686021.

References:

[1] H. Suzuki, T. Moriwaki, Y. Yamamoto, Y.Goto, “Precision Cutting ofAspheriacl Ceramic Molds with Micro PCD Milling Tool”, CIRP Annals, Vol.56, No. 1, pp. 131-134, 2007.

[2] Chunxiang Ma, T. Shamoto, T. Morimoto, Lijian Wang, “Study of MachiningAccuracy in Ultrasonic Elliptical Vibration Cutting”, International Journal ofMachine Tools and Manufacture, Vol. 44, No. 12-13, pp. 1305-1310, 2004.

[3] Y. Altintas, “Direct Adaptive Control of End Milling Process”, InternationalJournal of Machine Tools and Manufacture, Vol. 34, No. 4, pp. 1305-1310, 1994.

[4] D. Kurihara, Y. Kakinuma, S. Katsura, “Cutting Force Control ApplyingSensorless Cutting Force Monitoring Method”, Journal of Advanced MechanicalDesign, Systems, and Manufacturing, Vol. 4, No. 5, pp.955-965, 2010.

Figure 5 Behaviour of each response

231


Ultra Precision Process Monitoring

C. Brecher1, D. Lindemann1, A. Merz1, C. Wenzel1 1Fraunhofer Institute for Production Technology IPT Germany

[email protected]

Abstract

In modern applications the needs for highly advanced components with very precise

contour accuracy and surface quality are rapidly increasing. To meet the demands for

enhanced products, production machines and processes have to be improved with

respect to not only the high aims for precision, but also issues of standardization and

quality management. In this context process monitoring becomes a key technology to

achieve a better understanding of ultra-precision machining and to enable process

continuity, quality management and documentation. At Fraunhofer IPT a precision

process monitoring system has been developed combining data acquisition, analysis,

storage and visualization into one integrated system. The in-process sensor data is

mapped onto a virtual tool path generated using the linear scale positions of the

machine tool within the process monitoring system. Monitoring the sensor data in 3D

with regard to time and location of its acquisition enables variable evaluation

possibilities far beyond the state-of-the-art time plot or spectrum analysis methods.

1 Introduction – Process Monitoring using Acoustic Emission Sensors

In the manufacturing of precision components many influences affect the quality of

the work piece. Tool wear or micro damages to the diamond tool, the cutting

parameters, the dynamic machine behavior, material variations and environmental

influences can lead to lower quality of the machined work piece or even to rejected

parts. Process monitoring, being well established in standard machine tools, is an

effective method to detect, analyze and overcome malfunctions resulting from the

aforementioned influences. To detect process effects such as tool wear in an ultra-

precision diamond machining process, which is characterized by low process forces

and high demands for surface quality, very sensitive sensor probes have to be used.

Acoustic emission (AE) sensor probes, integrated for instance into the tool holder, are

very sensitive and provide high resolution and high frequency metrology data [1].

232



State-of-the-art process monitoring systems plot the acquired sensor data versus time

or further process the data e.g. for spectral analysis. For data processing the sensor

signal is first being transformed into a matrix description e.g. by short-time Fourier

transform (STFT). Afterwards the matrix can be treated as an image containing

discrete data information and can be analyzed using image processing strategies such

as Law’s analysis [2] or Haralik’s textural image classification [3]. In a final step

statistical analysis can be performed to find characteristic patterns for the individual

process. These methods have the disadvantage of losing the information of local

assignment to the work piece geometry and therefore lack of evaluation possibilities.

2 New Precision Process Monitoring Approach

To enhance the evaluation opportunities of process monitoring data in ultra-precision

processes, further information – the tool path, measured with the linear scales of the

machine tool axes – can be added to the process monitoring system. With this

approach a sensor signal analysis not only time-based, but locally referred to the

geometry of the work piece is possible. Using the 3D tool path, measured during

machining, and superposing the AE sensor signal color-coded, which means plotting

the sensor along a virtual tool path on the work piece, a 4D metrology plot results.

Performing the data acquisition, data processing and data visualization in real-time

while the part is being machined, a very flexible process monitoring tool has been

developed at Fraunhofer IPT.

The system has been designed as a black box consisting of industrial standard

components (IPC, metrology boards, connectors) and acquires all data signal

synchronously at high sampling rates. After the acquisition various pre-processing

features are implemented. The individual signals can be filtered using common digital

filters. They further can be scaled and converted so that the electrical signals match

the mechanical and physical conditions. Before the original analysis and visualization

routines are applied, the initial sensor data can be saved into a HDF5 container file

including metadata to describe the process monitoring environment [4]. Figure 1

illustrates the hardware and software structure of the novel process monitoring

approach.

233


Hardware

Software

AcquisitionMetrology Processing

VisualizationAnalysis Kinematics

Axis positions

Acoustic emission

Force, acceleration

Hardware driver

Sampling

Synchronization

Scaling

Filtering

Save, load

FFT, statistics

Image processing

Pattern recognition

Turning, milling

5-axis machining

Robotics

2D, 3D, 4D, 5D

Time domain

Frequency domain

50 100150200250

-2000

0

2000

50 100150200250

-2000

0

2000

50 100150200250

-2000

0

2000

Figure 1: New process monitoring approach – hardware and software algorithms

3 Experimental Results

To validate the system various experimental tests have been performed. A bearing

roll has been provided with a reference groove to analyze the sensitivity and

capability of local assignment of the system. Figure 2 shows a picture and a 4D plot

of the bearing roll after the finishing cut. The impact at the edges (reference groove

and tool entering and exiting the work piece) can be clearly seen. Figure 3 displays

the according 2D and 3D color-coded plots.

Bearing Roll – Machined Part & AE Sensor Plot

Figure 2: Experimental results of a bearing roll process analysis

234


Reference grooveReference groove

3D color-coded plot of bearing roll 2D color-coded plot of lateral surface

Figure 3: 2D and 3D color-coded plot of a bearing roll

4 Summary and Outlook

At Fraunhofer IPT a novel precision process monitoring system has been developed

and tested that maps sensor data onto the tool path while a work piece is machined.

With this method a local reference between part geometry and AE sensor signals is

possible enabling evaluation strategies beyond state-of-the-art monitoring methods.

Further research work will focus on porting and integrating the system to various

machines and processes to prove its comparability.

References:

[1] Köhler, J.O.A.; Schäfer, C.; et al.: Ein körperschallbasiertes

Überwachungssystem für die Ultrapräzisionsfertigung. 34th annual conference for

acustics, Dresden, 2008, pp. 823-824 (German).

[2] Laws, K.: Textured Image Segmentation. Ph.D. Dissertation, University of

Southern California, January 1980

[3] Haralik, R.M.; Kelly, G.L.: Pattern Recognition with Measurement Space and

Spatial Clustering for Multiple Images. Proceedings of the IEEE, Vol. 57, No. 4,

April, 1969, pp. 654-665.

[4] The HDF Group: http://www.hdfgroup.org/HDF5/doc/

235


Analysis of Mutual Influences of Control, Feedback and

Servo Drive Systems for Ultra Precision Machining

C. Brecher1, D. Lindemann1, C. Wenzel1 1Fraunhofer Institute for Production Technology IPT, Germany

[email protected]

Abstract

The mechanical design of ultra-precision machine tools is very well understood

today. Detailed investigations on precision axes designs, dimensioning of bearings

and drives and overall machine concepts have built a broad basis for designing very

stiff and accurate state-of-the-art machine tools. Enhancements to further increase the

achievable form accuracy and surface quality and at the same time decrease cycle

times and error sensitivity can only be accomplished by innovative control and drive

systems. In contrast to mechanical machine design, control, servo drive and feedback

as well as their interactional behavior within a complex machine setup have not been

sufficiently analyzed yet. This applies especially to ultra-precision machining. At

Fraunhofer IPT a test bench has been developed to analyze machine controls, servo

drives and encoder and sensor systems with regard to an evaluation of capabilities of

their application in an ultra-precision lathe. This paper will give a summary of the

results of servo drive and linear encoder analysis including both a comparison of

individual components and an investigation on mutual interactions.

1 Introduction – Test environment to analyze precision control systems

Investigating on all components applied in closed loop controls, their individual

performance and simultaneously mutual disturbances and limitations within the

whole system can be identified. Focusing on hardware structures, software modules

and data processing structures, an overall statement concerning all aspects of modern

closed loop control systems can be elaborated. A test bench has been configured as an

ultra precision lathe to later validate the measured results by diamond turning an

optical part. The setup uses two air bearing ironless linear drives and an air bearing

spindle. A modular mounting grid allows for the flexible integration of external

metrology such as laser interferometers, laser vibrometers or acceleration sensors.

236



Thus, a complete investigation on all aspects of precision motion control can be

guaranteed. The test bench setup, shown in Figure 1, has been presented in detail at

the 11th euspen International Conference 2011 in Como [1].

Z-AxisX-Axis

C-Axis, Spindle

Dampers

Granite Base

Laservibrometer

Laserinterferometer

Linear Scales

Z-AxisX-Axis

C-Axis, Spindle

Dampers

Granite Base

Laservibrometer

Laserinterferometer

Linear Scales

3 Linear Scales

Laserinterferometer

Laservibrometer

Accelerat ion Sensors

Z-AxisX-Axis

C-Axis, Spindle

Dampers

Granite Base

Laservibrometer

Laserinterferometer

Linear Scales

Z-AxisX-Axis

C-Axis, Spindle

Dampers

Granite Base

Laservibrometer

Laserinterferometer

Linear Scales

3 Linear Scales

Laserinterferometer

Laservibrometer

Accelerat ion Sensors

Figure 1: Test bench set-up with integrated metrology

The performed measurements include the analysis of the position accuracy and

repeatability (step response) as well as the determination of the dynamic frequency

characteristics (stiffness/ compliance) of an air bearing axis. With respect to the

aforementioned measurements the tests have been performed under the variation of

linear scales (vendor, pitch, signal, sampling frequency, etc.) and servo drives

(vendor, switching or linear amplifiers, PWM (Pulse-Width Modulation) frequency,

control architecture, DC bus voltage, etc.). First results have been presented at the

12th euspen International Conference 2011 in Stockholm [2].

2 Comparison of linear scales

As one aspect of the linear scale analysis an axis stiffness measurement has been

performed. Under variation of scale pitch (between 250 nm and 20 µm) the air

bearing axis has been excited with a piezo actuator with 40 N and a white noise

signal. The position deviation has been measured with a capacitive sensor. The signal

has been evaluated under statistic repetition to create a plot of the axis frequency

behavior. Figure 2 shows the results at 4 kHz (dynamic behavior) and 40 Hz (static

237


behavior). It can be seen that neither the dynamic nor the static stiffness shows

significant dependence on the encoder pitch distance.

Figure 2: Comparison of scales – axis stiffness at 4 kHz / 40 Hz (variation of pitch)

3 Comparison of servo drives

Regarding the servo drive system comparison a step response test (10 steps, width

between 10 nm and 10 µm) has been performed to analyze the position accuracy at

standstill and in motion as well as the repeatability. Servo drives have been analyzed

with respect to the PWM frequency and the comparison between linear and switching

amplifiers. Figure 3 shows the results of the 20 nm and the 10 nm measurements and

draws a comparison between a switching amplifier (8 kHz PWM) and a linear

amplifier as well as a comparison between two switching amplifiers (16 kHz and

100 kHz PWM). The standstill noise of the linear amplifier achieves the best results

(about 2 nm), but switching amplifiers with high PWM frequency nearly reach the

same performance. The higher the PWM frequency, the better the position accuracy.

238


Figure 3: Comparison of drives – position accuracy (variation of topology / PWM)

4 Conclusions and Outlook

At the Fraunhofer IPT a test bench has been developed and setup to investigate on the

influences of control components on the performance of ultra-precision axis. First

results confirm that axis stiffness does not depend on the encoder pitch and digital

servo drives with high PWM frequencies can reach the performance of linear

amplifiers. Future work will focus on a more detailed analysis of linear scales with

regard to signal quality, interpolation, sampling frequency and auto calibration as

well as on various aspects of CNC controls such as NC cycle time, interpolation

frequency and method or setpoint communication strategy.

References:

[1] Brecher, C.; Lindemann, et al.: Analysis of Control and Servo Drive Systems for

the Application in Ultra Precision Machining. Proceedings of the euspen 11th

International Conference, Como, 2011, ISBN 978-0-9553082-9-1, pp. 303 - 306

[2] Brecher, C.; Lindemann, D.; Wenzel, C.: Influences of Control, Feedback and

Servo Drive Systems on Precision Machining. Proceedings of the euspen 12th

International Conference, Stockholm, 2012, ISBN 978-0-9566790-0-0, pp. 344 - 347

239


Determining the random measurement errors of a novel

moving-scale measurement system with nanometre

uncertainty

N.Bosmans1, J. Qian1, D. Reynaerts1 1KU Leuven, Department of Mechanical Engineering, Belgium

[email protected]

Abstract

This paper describes a setup with a low sensitivity to temperature variations for

determining the random measurement errors of a measurement system applying a

moving scale. This moving-scale system is developed for advanced equipment such

as ultra-precision machine tools and should operate with a measurement uncertainty

of 15 nm for a measurement length of 109 mm and temperature variations of 1°C.

Temperature drift is identified as the most contributing source of errors and therefore

should be accurately determined. A dedicated setup has been designed for this task.

1 Introduction

1.1 Measurement system with moving-scale

Measuring displacement of the stages in ultra-precision machines is mostly done by

linear encoders or laser interferometers. Linear encoders can generally not be

configured like laser interferometers in an arrangement such that Abbe-offset is

eliminated for a multi-DOF system, but they outperform laser interferometers in

terms of stability w.r.t. environmental changes [1]. Earlier work at KU Leuven has

proposed a moving-scale measurement system in a configuration compliant with the

Abbe principle [2]. A prototype has been designed and experiments have been

conducted on critical components. Previous research has indicated it is possible to

reach a 15-nm measurement uncertainty of a 1-DOF moving scale measurement

system with a measuring length of 109mm [3].

1.2 Error budget and scope of paper

The error budget for the 109-mm moving scale measurement system is shown in

Table 1. The measurement uncertainty consists of random and systematic

measurement errors. Systematic errors, such as scale errors and Abbe errors resulting

240



from repeatable error motion of the guides, will be determined and eliminated in the

future by calibration with a laser interferometer. To determine the random errors

however, laser interferometry is not convenient since it would require extremely

stable ambient conditions or even a vacuum along the measurement path of the laser.

Therefore, the random errors, including quasi-static temperature drift, are determined

in a separate setup. This paper describes the design of this setup and discusses the

results of some preliminary experiments.

Table 1: Error budget of 109-mm moving-scale measurement system.

Component Value [nm]

Random measurement errors (±2σ) 8

Temperature drift reading head

Difference in drift between capacitive sensor and weather station

Other temperature errors (scale expansion, scale carrier expansion, ...)

Difference in humidity drift between cap. sensor and weather station

5

5

3

2

Dynamic errors 3

Systematic measurement errors (±2σ) 13

Abbe error

Other Geometric errors (Cosine error, ...)

5

3

Linear scale calibration error (estimation)

Other measurement errors (non-linearity, ...)

10

5

Total (±2σ) 15

2 Measurement setup for random errors

2.1 Design concept

Figure 1 shows the layout of the

setup. It consists of a moving-

scale measurement system with

a linear scale and a capacitive

sensor located on a scale carrier.

The scale carrier is driven by a

linear motor. The capacitive

sensor measures the

displacement of a target surface on another linear-motor-driven slide. The first linear

motor is controlled in such a way that the gap measured by the capacitive sensor

remains constant. The displacement of the target surface is calculated based on the

Figure 1: Measurement setup for random errors

241


readings of the linear encoder and the capacitive sensor on the moving-scale

measurement system. In order to compare the measured values, the position of the

target surface is simultaneously monitored by a linear encoder that is located on a

separate scale carrier containing the target surface and is in line with the first scale.

This second linear encoder gives the true displacement of the target surface and this is

used to compare with the measured value of the moving-scale measurement system.

Two Zerodur® interface discs provide a thermal fixed point at the front surface of the

capacitive sensor for the moving linear scale and at the target surface for the

reference scale. Consequently, the virtual distance between a point on the moving

scale and a point on the reference scale should not vary with temperature changes.

The position of the linear scales is measured using three reading heads RH1, RH2 and

RH3. The reading heads are fixed to an aluminium metrology frame. Because the

reference scale is made out of Zerodur®, which has a near-zero coefficient of thermal

expansion, the measured displacement between the two reference scale reading heads

equals the thermal expansion of the metrology frame. Thereby we assume that the

thermal expansion is uniform in the measurement direction.

The random measurement errors indicated in Table 1 are equal to

.

There will be a significant contribution of the random errors of the reference scale

system included in these measurements since they consist of the same error

components as the moving scale system, but without the drift of the capacitive sensor.

The random errors of the reference scale system will amount to 6 nm, bringing the

total random measurement errors to 10 nm (±2σ).

2.2 Preliminary experiments

An important part of the thermal stability is attained by proper mounting of the

reading head, a Heidenhain LIP28R. Therefore, the reading head is bolted to a carrier

made of the same material as the case of the reading head. This carrier is then

kinematically mounted to the metrology frame by three ball-in-V mounts. The point

where the lines through the V-grooves intersect is the thermal centre, which is at the

same position along the measurement direction as the thermal centre of the internal

242


grating of the reading head. In this way, the measurement drift of the reading head

will only be dependent on the expansion of the metrology frame and not on the

expansion of the reading head itself. To check the position of this thermal centre, a

setup has been built up and experiments have been carried out. Figure 2 shows the

setup and the drift of the reading head with changing temperature. Since the thermal

centres of reading head and scale coincide in this setup, there is negligible

temperature dependent drift.

(a) (b)

Figure 2: Setup for verification of thermal centre (a) and measurement results (b)

3 Conclusion

The design and preliminary experiments on critical components of a measurement

setup for random errors in a moving-scale measurement system are presented. The

setup, which is currently being manufactured, shall verify if the random errors will

not exceed 10 nm (±2σ). Experiments have verified that the temperature drift is

negligible once the thermal centre of the reading head is properly defined.

4 Acknowledgements

This work is supported by a PhD grant from the Institute for the Promotion of

Innovation through Science and Technology in Flanders IWT/101447, and the EC

FP7 FoF collaborative project - “MIDEMMA” (Grant agreement no. 285614).

References:

[1] H. Kunzmann, T. Pfeifer, & J Flügge, Scales vs. Laser Interferometers

Performance and Comparison of Two Measuring Systems, CIRP Annals -

Manufacturing Technology, 42, pp. 753 - 767, 1993

[2] D. Hemschoote, P. Vleugels, J. Qian, H. Van Brussel, D. Reynaerts, „An Abbe-

compliant 3D-measurement concept based on linear scales‟, Euspen 4th

International Conference proceedings, pp. 336 – 337, 2004

[3] N. Bosmans, J. Qian, J. Piot, D. Reynaerts, Design of a precision measurement

system using moving linear scales, Euspen 12th International Conference

proceedings, Vol. 1, pp. 302-305, 2012

243


An Approach to the Optimal Observer Design with

Selectable Bandwidth

I. Furlan, M. Bianchi, M. Caminiti, G. Montù

University of Applied Sciences of Southern Switzerland, Manno, Switzerland

[email protected], [email protected], [email protected],

[email protected]

Abstract

It is well know that the Kalman filter performs the best possible state estimator for

processes affected by noise. A consequence of the optimality, is that the convergence

speed of the estimation error cannot be selected by the user, because it depends by the

covariance of the noises. Since for several application this characteristic could be

undesired, this paper introduces an alternative design approach, in which the state

estimation error could be optimized for a given bandwidth by the user. The

effectiveness of the method is shown with an illustrative example.

1 Introduction

The classical full order observer for a linear system (A,B,C,D) is defined as follows

dxs/dt = Axs(t) + Bu(t) +L(y(t)-Cxs(t)-Du(t)) (Equation 1)

where: xs are the estimations of the real states x, u and y are the input and the output

of the system respectively, and the matrix L weights the correction given by the

deviation between the estimated output by the observer and the real output of the

system. Usually the two following methods can be used to determine the matrix L: the

matrix can be selected in order to obtain a desired convergence speed of the

estimation error xs-x or, alternatively, the matrix L can be determined in order to

minimize the variance of the estimation error E[(xs-x)2] by using the well know

Kalman-Bucy filter described in the seminal papers [1]. The first way allow the user

to decide the bandwidth of the estimation, but does not guarantees the optimality of

the estimation in presence of noises in the observed systems, counter-wise , the

second way, guarantee the optimality of the estimation but does not directly allow the

user to select the observer bandwidth. This paper introduces a third simple approach

244






to the observer design, in which for a given bandwidth decided by the user, the

optimal observer estimator is derived.

2 Alternative observer design method

Firstly the method will be introduced for multi output system, the case for single

output systems will be discussed later. In the case in which the system possesses m>1

outputs, the L matrix results of dimensions n x m. It is well known, if the system is

observable, that to place the n observer poles in a desired position, n free parameters

are required. Consequently, the matrix L performs (n x m)-n supplementary degree of

freedom that can be used for other purposes. The main idea presented by this paper,

consists in the use of the supplementary degree of freedom performed by matrix L, to

minimize the variance of the state estimation. More formally speaking, the following

mathematical problem has to be solved

minL E[(x-xs)2] subject to det(λiI-A+LC)=0 for all i ϵ{1,n},

where λi are the desired observer poles and n is the order of the system. A possible

brute-force search solution to the problem above will be introduced in the illustrative

example in the next paragraph; a more elegant solution to the problem will be subject

of further papers. For single output systems, the number of parameters is insufficient

to contemporaneously place the observer poles at a given position and simultaneously

minimize the variance of the estimation error, in this case the problem can be solved

by relaxing some conditions on the observer poles. An example could be the

following: impose the pole Euclidian norm of the poles but not the angle, obtaining

consequently, some free parameters in the matrix L that can be used to minimize the

variance of the estimation error.

3 Illustrative example

Problem set-up:

The state vector x(t)=[x(t),v(t)] of a classic mass damper system in figure 1 has to be

estimated.

figure 1

245


The dynamic equation is

ma(t)= F(t) – kx(t) – dv(t)

here a(t), x(t) and v(t) represent the acceleration, the position and the speed of the

mass m. F(t) is the input force, k and d are the spring and damping factor,

respectively. The parameters of the system are:

m=0.056 kg, k= 840 N/m, d=0.18 Ns/m.

The input force amplitude F(t) is known and the measurement of the position and of

the acceleration are available, i.e. y(t)=[x(t), a(t)]T. The input signal and the

measurements are affected by white Gaussian noises with covariance matrices

CU=4.47x10-8 N2 and CY=diag([5.08x10-16 m2,.4.47 x10-6 m2/s4]).

For control purposes, the bandwidth of the observer has to be 600 rad/s.

Solution:

Since the bandwidth has to be 600 rad/s, the following observer poles can be selected:

λ1 =-548 rad/s, λ2 =-670 rad/s. To determine the matrix L according to the principle

introduced by this paper, several procedures can be used; in this case the following

method has been applied. Since the order of the system is 2 and the output of the

system are 2, the matrix L result to be square composed by 4 elements, i.e.

L = [l11 , l12 ; l21 , l22].

In order to reduce the problem complexity, the task has been solved as follows. Since

one of the measurements (the acceleration a), corresponds to a derivative of a state

variable (the speed v), the parameters of L have been used to express the estimated

acceleration, i.e. the second element of the vector dxs/dt, as a weighted sum of the

acceleration determined using the model of the system called amodel(t), and the

measured acceleration a(t), i.e.:

estimated acceleration = (1- w). amodel(t) + w.a(t),

where w is the weighting factor, i.e. a real number in the interval [0,1]. With this set-

up, the coefficients l12 and l22 becomes 0 and w respectively, reducing by one the

number of parameters for the covariance minimization. The obtained observer

respecting the upper condition is:

dxs/dt = (I-Λ) (A-LC)xs(t) + (I-Λ).B.F(t) +L.y(t)+Λ.a(t)

where Λ=diag([1,l22]). Then, the optimal value of the parameter l22 has been obtained

by using a brute-force search in the interval [0,1], meanwhile the parameters l11 and

l21 have been determined to place the poles in the desired position.

246


The illustrative example has been solved by using also the standard methods, i.e.: the

pole placement method of Matlab, and by using the Kalman filter method [1]. All the

results have been reported in table 1. The outcomes show that the Kalman filter

version performs an observer with the better performance in sense of estimation

noise, but with a bandwidth lower than the desired one. Contrariwise the pole

placement algorithm of Matlab, performs a solution with the desired bandwidth but

with a not optimized noise. Instead, the suggested concept performs a compromise

between the two classical solutions.

Method Obtained bandwidth Estimation RMS noise

Standard pole placement 600 rad/s 5.63 x 10-4

Kalman 285 rad/s 1.72 x 10-8

Proposed method 600 rad/s 3.36 x 10-8

Table 1 : results

The proposed procedure can be generally applied when some measurements

correspond to the derivative or some states. We are working for a more general and

elegant method to determine the matrix L, according to the idea proposed in this

paper. This topic is the subject of our current researches.

4 Conclusions

The presented observer design method allows an optimal observer design in presence

of bandwidth constrains.

References:

[1] R. E. Kalman and R. S. Bucy, “New results in linear filtering and prediction

theory,” Trans. ASME—J. Basic Eng., vol. 83, pp. 95–108, 1961, ser. D.

247


Bandwidth Increase for Plate-like Structures by Adding

Mechanical Dampers

C.A.M. Verbaan1, P.C.J.N. Rosielle1, M. Steinbuch1

1 Control Systems Technology group, Department of Mechanical Engineering,

Eindhoven University of Technology, The Netherlands

[email protected]

Abstract

Precision designs often lack damping, which makes it difficult to achieve high

performance and sufficient robustness. This paper presents a method to increase the

open-loop cross-over frequency of a control system of positioning stages with a large

width-height-ratio by adding mechanical dampers to the system. Results are a smaller

tracking error (a factor 2 faster response) and better low-frequent disturbance

suppression.

1 Introduction

High-end processing steps often take place on positioning tables. These positioning

tables are part of high-tech motion systems with conflicting requirements like high

accelerations and accuracies in the sub-nanometer range. High frequent dynamics are

present in the mechanics and this often limits the achievable open-loop cross-over

frequency (bandwidth) of the motion system.1 Creating lightweight positioning table

designs with high natural frequencies is a classical way to enable high bandwidths.2

The damping in precision designs is generally low, which leads to large amplification

magnitudes at resonance frequencies. Monolithic designs, ceramic materials and

vacuum operating environments contribute to this lack of damping. Tuned mass

dampers (TMD) are commonly used in dynamic structures to dissipate energy and

suppress the vibration amplitude at a specific resonance frequency.3,4 In this research,

a damper with a comparable structure is used to add damping to a range of high

frequent resonances, which leads to a substantial increase in bandwidth of the

positioning table.

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2 Undamped stage behaviour and limits

An unconstrained square plate is investigated as abstracted model of a positioning

stage. This abstraction is justified by the often relatively large width-height-ratio of

these tables. An undamped modal analysis with shell elements is performed using the

finite element method (FEM), and a number of mode shapes and corresponding

natural frequencies are exported, including rigid body modes. The first mode shapes

of flat structures are characterised by large out-of-plane displacements. Therefore, the

behaviour in z-direction is studied. A state space description of the plate in modal-1

form5 is generated and a modal damping of 10-3 [-] is assumed to be present. The

plate is provided with position sensors on the corners (4x in z-direction) and force

actuators (4x in z-direction) that are located 0,2l from the plate sides with length l.

See figure 1a. A geometrical decoupling procedure with respect to the centre of

gravity (CoG) of the plate is applied to extract the transfer function in z-direction.

The decoupled transfer function in z-direction, w.r.t. the centre of gravity, is shown in

figure 1b. A controller has been designed for the plant shown in figure 1b. The

controller consists of a gain, lead-filter and 1st order low-pass filter and usual

robustness margins are applied (Mod. mar: 6 dB / Ph. mar: 30 deg / Gain mar: 6 dB).

The achieved bandwidth is 28 [Hz]. In this case, the third resonance appearing is

limiting the loop gain. This resonance lacks a preceding zero. Therefore the phase

decreases from -180 deg to -360 deg. The pole of the low-pass filter has to be placed

at relatively low frequencies to deal with this phenomenon.

(a)(b)

Figure 1a: The undamped plate model with four actuators and four sensors in z-direction

Figure 1b: Transfer function in z-direction, decoupled around CoG

102

103

-200

-180

-160

-140

-120

-100

Plant bode diagram

Mag

nit

ud

e[m

/N]

102

103

-360-270-180

-900

Ph

ase

[deg

]

Frequency [Hz]

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3 Damped stage behaviour

To increase the structural damping, mass-spring-damper systems are added to the

plate corners. See figure 2a. The DoF of the masses are in z-direction and the mass

contribution of the dampers is 0.35% per damper w.r.t the plate mass. For most mode

shapes, the plate corners have relatively large displacements. Therefore, also

velocities are relatively large during vibration. Energy is extracted from the plate by

the mass-spring-damper systems, which results in damped behaviour of the

resonances. The bandwidth limiting resonances are suppressed by adjusting the

natural frequency and damping of the damper systems. See figure 2b. For comparison

reasons, a controller of the same order (number of controller components) and same

robustness margins has been designed for the damped plate as in case of the

undamped plate. The bandwidth achieved equals 74 Hz, which is 2.5 times higher

than the bandwidth corresponding to the undamped plate. Figure 3 shows both open-

loops in a Bode diagram. The controller gain in case of the damped plate is increased

substantially with respect to the undamped plate, as a result of the damped

resonances. In addition, the pole(s) of the low-pass filter can be placed at higher

frequencies, because a major part of the occurring resonances is damped. This causes

less phase lag at the bandwidth and therefore it is easier to preserve the stability

margins and increase the bandwidth.

4 Conclusions

This paper shows a robust way to add damping to relatively flat positioning stages.

The damping reduces the amplitudes at resonance frequencies, through which the

Figure 2a: The damped plate model with four dampers added in z-direction

Figure 2b: Transfer function in z-direction, decoupled and with dampers added.

102

103

-200

-180

-160

-140

-120

-100

Mag

nit

ude

[m/N

]Plant bode diagram

102

103

-360-270-180

-900

Pha

se[d

eg]

Frequency [Hz]

Undamped plate

Plate with 4 TMD's

(b)(a)

250


controller gain can be increased. The results are a substantial bandwidth increase

which leads to faster responses to setpoints, smaller tracking errors and a better low-

frequent disturbance suppression. In general, this approach provides additional design

parameters to the mechanical engineer during the design phase of a positioning table.

The combination of a precision mechanical design extended with dampers that are

specifically designed for this positioning table leads to substantial performance

increase of the overall system.

AcknowledgementThis project is supported by ASML Research Mechatronics (NL). We greatly

appreciate the input of dr.ir. J.P.M.B. Vermeulen, dr.ir. M.M.J. van de Wal and dr.ir.

S.H. van der Meulen.

References:[1] Skogestad & Postlethwaite, Multivariable Feedback Control, John Wiley & SonsLtd, Chichester, 2005[2] Book, W.J., Controlled Motion in an Elastic World, Journal of DynamicSystems, Measurement and Control, 50th Anniversary Issue, March 1993, pp 252-261[3] J.P. Den Hartog, Mechanical Vibrations, McGraw-Hill, New York, 1956[4] S. Krenk, J. Hogsberg, Tuned mass absorbers on damped structures, 7th

European Conference on Structural dynamics, July 2008[5] W.K. Gawronski, Dynamics and Control of Structures – a modal approach,Springer-Verlag, New York, 1998

102

103

-80

-60

-40

-20

0

20

Mag

nit

ude

[dB

]

Bode diagram Open-loop

Undamped plate

Plate with dampers

102

103

-540

-360

-180

0

Phas

e[d

eg]

Frequency [Hz]

Figure 3: Bode diagrams of open loop. The dotted line represents the undamped plate and the

solid line represents the plate with the dampers. The bandwidth increase is indicated.

251


A Parallelism Alignment Mechanism for Nanoimprint

Lithograph with Large Imprinting Force

W.J. Chen, W. Lin, G.L. Yang

Singapore Institute of Manufacturing Technology (SIMTech), Singapore

[email protected]

Abstract

This paper proposes a force-bypassed parallelism alignment mechanism to address

the negative effect of the imprinting force in Nanoimprint Lithograph (NIL). It

enables the imprinting force bypass the delicate compliant members, thus ensuing the

active and passive parallelism alignment able to be carried out properly even under a

large imprinting force. A prototype of the parallelism alignment mechanism has been

developed and tested. Experimental results show that superior alignment accuracy

still can be achieved under an imprinting force up to 1KN.

1 Introduction

Nanoimprint Lithograph (NIL) utilizes the imprinting force to transfer circuit patterns

from a template to a substrate. Its two critical specifications i.e. overlay accuracy and

pattern transfer fidelity, depend on performance of the parallelism alignment system.

Parallelism alignment in NIL needs to perform out-of-plane (θx and θy) motions,

which brings the template and substrate surface into parallel contact while

minimizing the lateral motion during the imprinting process. This is typically

implemented through adopting an active alignment to eliminate large wedge errors

firstly and then a passive alignment to compensate the residual errors [1]. The active

alignment is done by a motorized precision stage, while the passive one done by

virtue of the deformation of a compliant mechanism under imprinting force.

Imprinting forces may vary from a few Newtons to several hundred Newtons in

different tasks. A large imprinting force will deteriorate the alignment performance if

the compliant mechanisms directly undertake such a force.

To compensate the wedge error with a relative higher sensitivity, most of off-the-

shelf complaint mechanisms [2] used for NIL parallelism alignment are arranged

252



within the force loop of compression. This configuration is acceptable only for cases

with small imprint forces. For cases with large imprint forces, such a design may

result in the compliant mechanism a significant deformation in unwanted directions,

consequently decreasing the overlay accuracy between the template and the substrate.

This paper will address this issue through presenting an imprinting force bypassed

parallelism alignment mechanism.

2 Descriptions of parallelism alignment mechanism structure

Figure 1 is an assembly view of the NIL press head developed in SIMTech. It

includes a template unit and a parallelism alignment unit. The template unit

comprises a template, a heating block, isolate plates and three force sensors, etc. The

alignment unit carries the template unit to perform required parallelism alignment.

Actuator-1

Bracket

Spherical

joint cup

Temperate

Actuator-2

Spherical

joint hump

Heater

block

Force

sensor

Heat

Isolate

plate

Compliant

mechanism

Supporting

plateCollar

Spherical

joint humpSpherical

joint cup

Bracket

Flexure

Adaptor Hump

shaft

Figure 1: NIL press head Figure 2: Structure of parallelism alignment unit

Shown in Figure 2 is the structure of the parallelism alignment unit, which comprises

a special spheroidal joint, a compliant mechanism, and two actuators. The spheroidal

joint is a reverse ball-and-socket joint, consisting of a ball-shaped hump and a thin

socket part. The hump can rotate about the center of the template surface. Unlike

common ball-and-socket joints, the shaft here is fixed on the top of the hump and

goes through the top opening of the socket. The compliant mechanism is anchored on

the top of the socket (or cup). Its platform (central portion) is firmly connected with

the hump shaft. The stiffness of the platform after considering actuators connection is

shown in Table 1. Obviously, the platform allows the hump θx and θy tilting motions,

but limits the hump θz motion due to the high stiffness value.

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Table 1: Stiffness of compliant mechanism platform

Translational Stiffness (N/μm) Angular Stiffness (N•m/deg)

Kx Ky Kz Kθx Kθy Kθz

25.5 25.5 4.7 22.0 22.0 395.5

3 Function analysis of parallelism alignment mechanism

A schematic of working principle of the parallelism alignment unit is shown in Figure

3, in which the non-uniform press forces distributing on the template is compounded

as a linear imprinting force F and a small torque M at the template centre. It can be

seen that imprinting force F goes through the template block, spherical joint hump,

spherical joint cup and bracket, finally, reaching to the machine frame. In other

words, the compliant mechanism does not bear the imprinting force F, but only

undertake the small torque M during compression. This feature means the large

imprinting force will not affect the behaviors of the compliant mechanism.

Figure 3: Working schematic Figure 4: Active/passive alignment in one setup

With the force-bypassed feature, the press head allows both active alignment and

passive alignment tasks in one setup. Before the template is brought to contact with

the substrate, an active alignment task is conducted to eliminate the coarse tilting

errors (Figure 4 (a)). Two actuators drive the complaint mechanism to carry the hump

to perform tilting motions about the template surface center until the required parallel

accuracy is reached. After the template contacts with the substrate, a passive

alignment is applied to eliminate the residual error of the active alignment. As shown

in Figure 4 (b), under a non-parallel contact, the press force on the template will

offset from the rotation center. This offset will result in the template block self-

rotating about the template center to eliminate the non-parallelism.

254



Press force uniformity is an important indicator of the press head parallelism. Three

force sensors uniformly installed on the template unit are used to measure and

monitor the press force distribution. The total press force of about 1KN is gradually

applied on the template and then release to zero. Figure 5 (a) shows the distribution

of the press force before alignment. The maximum variation of the press force over

the template is over 20 percent. Figure 5 (b) shows the press force distribution after

alignment. It can be found that the variation of press force is less than 5 percent.

Under such force uniformity, the parallelism error over the template area (φ50mm)

can be less than 20 nm.

Force Sensor Reading

-50

0

50

100

150

200

0 50 100 150

Time (s)

Forc

e (N

)

Sensor1

Sensor2

Sensor3

Force Sensor Reading

-50

0

50

100

150

200

250

300

350

400

0 50 100 150 200 250 300 350

Time (s)

Forc

e (N

)

Sensor1

Sensor2

Sensor3

(a) (b)

Figure 5: Force sensor reading (a) before alignment; (b) after alignment

5 Conclusions

Large imprinting force may affect the alignment accuracy in NIL. Using the “smart”

mechanism design presented in the paper will reduce the negative effect of the

imprint force.

References:

[1] Byung Jin Choi, Sidlgata V. Sreenivasan, Stephen C. Johnson, “High precision

orientation alignment and gap control stages for imprint lithography process”, US

patent No: 6873 087B1, Mar 29, 2005.

[2] Jae Jong Lee Kee-Bong Choi and Gee-Hong Kim “Design and analysis of the

single-step nanoimprinting lithography equipment for sub-100 nm linewidth”

Current Applied Physics Volume 6, Issue 6, October 2006, pp. 1007-1011

255


Design and Performance of a 6 DOF Hybrid Hexapod

N.L. Brown1, C.W. Hennessey1 1ALIO Industries, USA

[email protected]

Abstract

Hexapods, also known as parallel kinematic machines or Stewart platforms (see

Figure 1a and 1b), have no moving cables, increased dynamic response, smaller

moving mass, better Z stiffness, and smaller size compared to serial kinematic

systems making them the accepted six axis motion system for precision applications

[1,2,3,4]. As precision motion requirements increase from the 10’s of micrometers

to the nanometer level, existing hexapods cannot meet motion system requirements

due to performance limitations inherent in existing designs. A new six degrees of

freedom hybrid parallel and serial kinematic design (see Figure 1c) is presented that

addresses the performance limitations of hexapods and achieves sub-micron

accuracy, repeatability and straightness as required for advancements in optical,

semiconductor, manufacturing, metrology, and micro-machining industries.

(a) (b) (c)

Figure 1: 6 DOF motion systems: a) ALIO HR2 hexapod, b) example of a six axis

parallel kinematic layout, and c) hybrid parallel and serial XY- Tripod-Theta system.

1 Hexapod Limitations

Hexapods have six links joined together moving a common platform and thus the

motion error of the platform will be a function of the errors of all links and joints.

Hexapods are known to have optimum accuracy, repeatability, and path integrity

when performing Z axis moves because all links perform the same motion at the

same link angle. When any other X, Y, pitch, yaw, or roll motion is commanded,

accuracy and geometric path performance of the hexapod degrades because all links

256



are performing different motions [5]. Additionally, hexapods are marketed as

having good stiffness compared to serial stacked multi-axis systems [2,3,4,5].

However it is only the Z (vertical) stiffness that is acceptable. Stiffness has a large

impact on platform repeatability and rigidity and thus the relatively weak XY

stiffness, which is 10-30 times less than Z stiffness (Table 1), negatively affects XY

axis performance. Lastly while there are documented compensation methods to

reduce error sources, they do not improve performance to the sub-micron level.

2 Design Summary: Hybrid 6 DOF System

The presented hybrid parallel and serial kinematic system is designed to address and

minimize link and system error sources to achieve nanometer order performance.

The system is a serial stack of an XY stage, a redesigned parallel kinematic tripod

(Z, pitch, and roll), and a rotation (yaw) stage, see Figure 1c. The tripod includes a

new link design with precision linear crossed-roller bearings, non-contact optical

linear encoders, and brushless linear servo motors oriented along the link axis. This

design eliminates backlash, micrometer screw pitch errors, and error sources from

rotary encoders. Near frictionless pneumatic or magnetic counterbalances in each

link maintain high payload capabilities. This redesigned tripod is joined with XY

and rotary stages that provide optimized performance for XY and yaw motions. In

this hybrid concept, individual axes can be customized to provide travel ranges

ranging from millimeters to over one meter and still maintain nanometer levels of

precision. The following sections describe the improvements of this hybrid structure

in reference to the specific performance capabilities.

2.1 Linear Displacement Accuracy and Repeatability

Hexapod linear accuracy is limited to micron order performance that varies greatly

throughout its range of travel. The hybrid six DOF system pairs the optimized Z

axis performance of a parallel kinematic tripod with an XY stage calibrated to have

sub-micron accuracy and straightness performance. XY error sources are reduced

relative to the hexapod resulting in accuracy error less than +/- 2um in Z and less

than +/- 1 um in X and Y. The repeatability of a hexapod depends on the

repeatability in three dimensional space of all six links while the hybrid design’s

257


improved links combined with high precision serial stages enables repeatability

below +/- 50nm.

-4-3-2-101234

-100 -50 0 50 100LIN

EA

R D

ISP

LAC

EM

EN

T

AC

CU

RA

CY

ER

RO

R (

um

)

TRAVEL (MILLIMETERS)

XYZ AXIS - LINEAR ACCURACYX Axis

Y Axis

Z Axis

0255075

100125150175200

0 4 8 12 16 20

PO

SIT

ION

(N

AN

OM

ET

ER

S)

CYCLE

XYZ REPEATABILTY

X Axis Y Axis Z Axis

(a) (b)

Figure 2. The X, Y, and Z linear accuracy (a) and repeatability (b) of the hybrid six

DOF system, model AI-6D-300XY-106Z-104R, tested per ASME B5.54:2005.

Note: In plot (b) axis repeatability data is offset vertically for clear visualization.

2.2 Motion Trajectory: Straightness

Hexapod straightness of travel is often not quantified by manufacturers because it

can be greater than 100um/100mm travel, which is the result of parasitic errors from

all six links. With a precision XY stage the link error sources affecting path integrity

are reduced to error sources from two XY axes for which the performance can be

tightly controlled. Hybrid system straightness is less then +/-2um/100mm of linear

travel, which is two orders of magnitude better than typical hexapods, see Figure 3.

-3

-2

-1

0

1

2

3

-100 -50 0 50 100

ST

RA

IGH

TN

ES

S E

RR

OR

(u

m)


XYZ AXIS - STRAIGHTNESS X Axis

Y Axis

Z Axis

-3

-2

-1

0

1

2

3

-100 -50 0 50 100

ST

RA

IGH

TN

ES

S E

RR

OR

(u

m)


XYZ AXIS - FLATNESS X Axis

Y Axis

Z Axis

Figure 3. The X, Y, and Z straightness and vertical straightness (flatness)

performance of the hybrid six DOF system, model AI-6D-300XY-106Z-104R,

tested per ASME B5.54:2005.

2.3 XY Stiffness

In the hybrid system structure the parallel kinematic tripod link lower joint to the

base plate has only one degree of freedom and thus the combination of three joints

258


provides mechanical stiffness in the XY directions. Additionally, the XY stage

motors are oriented in the XY directions such that the full motor force determines

the XY servo stiffness. These

changes increase the XY to Z

stiffness ratio to 1:1.2, see

Table 1. The hybrid six DOF

system is ideal for machining,

optical, and laser manufacturing

applications where electrostatic,

bonding, or mechanical forces

are applied to the motion

system.

3 Conclusions

There are inherent weaknesses of the hexapod concept that prohibit the use of

hexapods in applications requiring sub-nanometer precision. The hybrid parallel and

serial kinematic system presented takes advantage of the Z, pitch, and roll capabilities

of parallel kinematic tripod and uses serial kinematic stages to provide the X, Y, and

yaw axes. The result is a six DOF motion system that can meet increasing motion

system needs for sub-micron performance.

References:

[1] Merlet J.P. Parallel Robots (Solid Mechanics and Its Applications). Springer.

New York: 2006.

[2] Parallel Kinematics Motion Systems, http://www.alioindustries.com.

[3] Hexapods (Stewart Platforms) Overview, http://www.physikinstrumente.com/.

[4] Robotics, Hexapod, www.pimicos.com.

[5] S. Szatmari, “Kinematic Calibration of Parallel Kinematic Machines on the

Example of the Hexapod of Simple Design”, Dissertation, Dresden University of

Technology, 2007.

[6] 6-Axis Parallel Kinematic Positioning System, www.newport.com.

Table 1. Published specifications for hexapods

and the hybrid 6 DOF motion system showing

the differences in XY to Z stiffness.

Manufacturer ModelRatio

XY : Z StiffnessType

M-850 1 : 33

H-850 1 : 14

H-824 1 : 4

HXP50-MECA 1 : 11

HXP1000-MECA 1 : 10

AI-HEX-HR2-SS 1 : 10

AI-HEX-HR4 1 : 14

AI-6D-100XY-24Z-56R 1 : 1.2

AI-6D-300XY-55Z-104R 1 : 1.1

Published Stiffness Specifications:

Hexapods vs ALIO Hybrid 6 DOF

ALIO Industries [2]

ALIO Industries [2]

Newport [6]

Physik Instrumente [3]

Hybrid 6 DOF

System

Hexapod

259

http://www.alioindustries.com/

http://www.physikinstrumente.com/

http://www.pimicos.com/

http://www.newport.com/


Concept design of a 5-axis portable milling machine for the

in-situ processing of large pieces

J. Eguia1, O. Gonzalo1, M. San Martín1, S. Ilhenfeldt2 1IK4 - TEKNIKER, Spain 2Fraunhofer IWU – Germany

[email protected]

Abstract

For conventional machine tools the work piece is placed inside the structure and

therefore its overall dimensions determine the size of the workspace and with it also

the general size of the complete system, which leads to extreme disproportions

between theoretically appropriate and actually needed system size. In large pieces this

results in even larger machines presenting technical and practical issues well known

and described in the state-of-the-art. An alternative to solve this issue is to get rid of

the dogma “work piece inside the machine” and replace it by the principle “small

machines on large work pieces” [1][2].

This paper presents and discusses a new paradigm of portability and proposes the

basic design of an advanced portable machine for in-situ and on-the-part milling of

large components. More specifically, the machine shall be able to machine features

larger than itself without continuity issues and automatically by means of an array of

sensors and built-in CAM reprocessing capabilities, in close contact with the CAD

file of the piece. For enhanced functionality, the machine targets usual engineering

materials (steel, aluminium and composites) and processes (drilling and milling) with

mid-to-high removal rates with a fully-flexible 5-axis configuration. Said machine

will become the backbone of a set of R&D efforts in the field of miniature and

portable machines with a view to developing sound solutions for the limitations of

portable machines in terms of part and feature precision, machine clamping on the

part and process capabilities with limited machine sizes.

1 Background and portability as a new paradigm

The basic idea behind the new paradigm proposed in this paper is to use autonomous

machining units which are placed locally at the work piece using it as machine

260



foundation. The dimension of these units and their performance no longer depend on

overall dimensions but on the geometry which is to be machined. This approach

predominantly follows the principles of miniaturisation, freedom through mobility

(whether it moves by itself or is displaced) and flexibility (in processes and

materials), and utilisation of synergies.

Moreover, this “small machine on large work piece” approach leads to improvements

in the followings aspects of the machine and process:

Mobility: meaning both the general transportability of production systems to the site

of operation and close to the work pieces, and the placement of the machinery on/at

the work piece. Miniaturization: a general reduction of system size. Adaptivity:

defined as a short term modulation of machine properties to match production process

requirements with the capabilities of the executing system. Mutability: defined as the

long term modifications on production systems. Multifunctionality: especially work

pieces with complex geometric features can be machined much faster and more

efficient through complete machining on multifunctional machine tools because of

elimination of work piece transport between different work stations. Specialisation:

specialised production technology may be provided with process fitted equipment,

thus ensuring very effective and efficient application of operating devices.

2 The fully-portable, five-axis, miniature milling machine

2.1 Machine concept

The machine proposed to establish the limits of this new paradigm is a five-axis,

miniature milling machine based on a serial kinematic architecture. This kinematic

solution has been chosen because it shows remarkably homogeneous stiffness

behaviour for every possible machine orientation and process combination. The

ability to easily configure and control the work volume and the simplified error

control and calibration procedures also contribute to favour this kinematic over the

parallel-machine architecture, even if the latter can show better stiffness-weight

ratios.

Over the state of the art [3] [4] [5] , this portable machine can perform both mid-duty

milling and drilling operations in a five axis configuration. To achieve this, three

stacked linear axes are included, which carry a two axis rotating system holding the

spindle, in a compact machine envelope of 1200 x 1200 x 1200 mm. The work

261


volume of the machine is 340 x 300 x 220 mm with A and C axes ranging from -40º

to 100º and 0 to 420º, respectively.

Figure 1: The concept design of the machine (man portrait for scale purposes)

2.1 The spindle

The machine is equipped with a 9 [email protected] rpm synchronous spindle which

allows the machine to cover mid-duty roughing operations in conventional steel (at

low speed and high torque) and operations in aluminium components (at higher

speeds and high power). With the two axes system right behind directly holding the

spindle, the process flexibility is therefore ensured.

2.2 Sensors

This paradigm makes the machine closer to a free robot than a conventional machine

and thus new needs appear for external referencing (navigation) and internal

referencing (feature recognition within the workpiece).

For internal referencing, the machine is equipped with a laser scanner with a typical

precision of 50 µm, combined with a touch probe for redundancy and finer data

acquisition.

For navigation, the machine relies ultimately in a laser tracker to define its position in

space although several more economic technologies are being studied for smoother

integration in industrial environments, such as the identification of scanned-part

features in the cad, or the use of bespoke fiducials.

3 Advanced process capabilities

The machine is intended to perform advanced manufacturing according to three

different scenarios.

262


Figure 2: The three machining scenarios considered

For scenarios 2 & 3, the machine uses both global and local referencing capabilities

in the following way:

Step1: Fix the relative position (machine – workpiece)

Step2: Scan the workpiece zone within the machine work space

Step3: Identification of the workpiece comparing the 3D data to the CAD file, using

singular features or previously located patterns for fitting

Step4: Definition of a reference coordinate system related to the machine axis

Step5: Definition of the tool path for the current machine position (CAM module)

Step6: Perform the machining operation

Step7: Back to first step

4 Conclusions

This paper presents a fully-flexible 5-axis miniature milling machine as a platform to

test the functional and productive capabilities of this novel portability paradigm. The

theoretical capability assessment and the experimental validation will be performed in

late 2013 and 2014.

References:

[1] Allen, et al. 2010, A review of recent developments in the design of special

purpose machine tools with a view to identification of solutions for portable in-situ

machining systems, I. J. Adv. Manuf. Tech., 50/9-12, 843-857

[2] Allen J.M., et al. 2012Free leg hexapod Proceedings of the Institution of

Mechanical Engineers. Part B, Journal of Engineering Manufacture. 226(3), 412-

430.

[3] ElectroImpact http://www.electroimpact.com/Flextrack

[4] Mirage portable machine tools, http://www.miragemachines.com

[5] York portable machines, http://www.yorkmachine.com

263

http://www.miragemachines.com/

http://www.yorkmachine.com/


Using Boron Doped Diamond Foils for Fabrication of Micro

Cavities with EDM

E. Uhlmann1, M. Langmack1, J. Fecher2, S. M. Rosiwal2, R. F. Singer2

1Institute for Machine Tools and Factory Management,Technische Universität Berlin,

Germany 2Institute of Science and Technology of Metals (WTM), University of Erlangen-

Nuremberg, Erlangen, Germany

[email protected]

Abstract

High precision cavities come into action for micro injection and micro embossing

tools in the field of tool making and are mainly used for small batch or mass

production of micro parts.

To fabricate a large quantity of parts, wear resistant tool materials are required.

Having a high hardness and a high Young´s Modulus, the materials used are often

heavy or even impossible to machine by conventional fabrication processes. Being

independent of the work piece’s mechanical properties Micro Electrical Discharge

Machining (µEDM) is predestined in this case.

Besides the adjustment of the electrical parameters, the µEDM-process is also

determined by the tool electrode`s material having a big influence on the machining

time, the result, and the electrode´s wear behaviour [1]. To assure an efficient

process, short production time and a low tool wear are demanded. Therefore,

electrodes with an excellent electrical and thermal conductivity as well as a high

mechanical strength have to be used.

1 Introduction

The investigations described in this article show experimental results concerning the

usage of two diamond foils with different boron dopings. Besides the examination of

their removal rate and the wear behaviour, the paper gives information about a

fabrication method concerning the manufacture of the foils.

264



2 Manufacturing of Diamond Foils by Hot-Filament Chemical Vapour

Deposition (HF-CVD)

For the manufacturing of diamond foils by HF-CVD a 3-step process is used, which

is described in detail by Lodes et al. [2]. Prior to diamond deposition the 6” silicon

wafers, which are used as substrate material, were seeded with a suspension of 4 nm

diamond powder and ethanol (1st step), which ensures a homogenous and fast

diamond growth directly from the beginning of the deposition process. The diamond

deposition (2nd step) was performed in a Cemecon CC800/Dia-9 HF-CVD plant in

hydrogen atmosphere containing 3.6 % methane in the gas flow at gas pressures of 6-

8 mbar. To obtain electrically conductive boron doped diamond foils, 0.5-1.5 %

trimethylborate was added as boron precursor gas in the gas flow. For HF-CVD-

diamond deposition tungsten filaments, which are electrically heated to 2.200 °C, are

used to enable the chemical reactions for diamond deposition and to heat the substrate

to 850 °C. The grain size of the diamond foils can either be adjusted by variation of

the gas pressure or by variation of the methane and boron contents in the gas phase.

The freestanding (i.e. substrate free) diamond foils were obtained by laser cutting and

an ultrasonic delamination treatment (3rd step) of the coated wafers. The boron

content of the boron doped diamond foils was measured by Glow Discharge Optical

Emission Spectroscopy (GDOES).

3 Process behaviour

For first experiments, boron doped diamond foils with a thickness of t = 30 µm were

used. Herewith micro cavities of a depth of dc = 150 µm and a width of wc = 50 µm

were fabricated in 5 mm probes made of 90 MnCrV 8 by µEDM. A no-load voltage

of u0 = 100 V at a discharge current of ie = 2.4 A was applied. Figure 1 shows the

machined cavity and a boron doped diamond foil after its use in the dielectric oil

IME 63. In general, the boron doped diamond foils showed a levelling of the B-CVD

diamond crystal on the contact area of the electrode.

Due to a high electrical field and a high thermal exposure during µEDM the B-CVD

diamond foils also showed an edge rounding. The machined cavities had straight even

side walls and an arithmetical mean deviation of the roughness profile of around

Ra = 0.3 µm at the cavity bottom. As a consequence it can be stated that B-CVD-

diamond foils were generally applicable as tool electrode material for µEDM.

265


Process: Working fluid:

Micro die sinking (µEDM) with stationary tool electrode Dielectric oil IME 63

Work piece electrode: Generator:

90MnCrV8 probe with bw = 5 mm Static Pulse Generator

ie = 2.4 A

Tool electrode: u0 = 100.0 V

BCVD-Diamond foil with bw = 30 µm t0 = 10.0 µs

200 µm 500 µm

50 µm20 µm

a) b)

Figure 1: a) Top view of a micro cavity and b) side view of a boron doped diamond

foil after machining

4 Removal rate and relative frontal wear

Further investigations focused on the analysis of the removal rate and the relative

frontal wear of B-CVD diamond foils having two different boron dopings, such as

0.23 at% and 0.3 at% (Figure 2). For experiments, probes made from 90MnCrV8

came into action. A static pulse generator was applied providing discharge currents of

ie = 2.4 A and ie = 3.2 A. Due to an increasing electrical conductivity, diamond foils

with a bigger boron content showed a higher removal rate then diamond foils with

lower boron content. Also, the resulting higher current flow caused an increase of the

relative wear. Therefore, an increase of the discharge current resulted in a growing

removal rate and a growing relative frontal wear for both foil electrode types.

5 Conclusion

First experimental investigations on the process and wear behaviour of B-CVD

diamond foils being used as tool electrodes for the µEDM process were presented.

The foil electrodes showed a general applicability when using static discharge pulses.

Within this paper, machining results when using different discharge currents and

diamond foils of different boron dopings were described.

266


0

15

30

%

60

rel. W

ear

rel

Foil 1(0.3 at%)

Type of BCVD Foil

Foil 2(0.23 at%)

Foil 2

(0.23 at%)

Foil 1(0.3 at%)

0,000

0,005

0,010

mm3/min

0,020

Foil 1(0.3 at%)

Rem

oval ra

te V

W

Type of BCVD Foil

Foil 2(0.23 at%)

Foil 2

(0.23 at%)

Foil 1(0.3 at%)

Process: Working fluid:

Micro die sinking (µEDM) Dielectric oil IME 63

with stationary tool electrode

Process Parameters:

Work piece electrode: Static Pulses

90MnCrV8 u0 = 100.0 V

probe with bw = 1 mm ti = 7.5 µs

t0 = 10.0 µs

Tool electrode:

BCVD-Diamond Discharge current ie = 2.4 A

foil with bw = 30 µm Discharge current ie = 3.2 A

a) b)

0.000

0.005

0.010

0.020

mm3/min

Figure 2: a) Removal rate VW and b) relative Wear ϑrel of diamond foils with different

boron dopings

Further investigations will focus on the optimisation of the B-CVD diamond foils.

For this purpose, further experiments describing the influence of the boron

concentration within the diamond foils on the process behaviour need to be carried

out.

6 Acknowledgments

The authors would like to thank the german research foundation (DFG) for

supporting this research.

References:

[1] Uhlmann, E.; Piltz, S., Roehner, M.: Influence of Diamond Coatings on

Electrode Wear in µEDM, Proc. of the 7th Int. Conference of the

European Soc. for Precision Engineering and Nanotechnology, Bremen,

Vol. II, pp. 525-528, 2007

[2] Matthias A. Lodes, Stefan M. Rosiwal, Robert F. Singer: Self-supporting

nanocrystalline diamond foils – a new concept for crystalline diamond on

any technical surface, Key Engineering Materials Vol. 438, pp. 163-169,

2010

267


Design and Optimization of Flexure-Based Micro-

manipulator for Optics Alignment

C. Brecher, N. Pyschny, T. BastuckFraunhofer Institute for Production Technology IPT, [email protected]

Abstract

This paper presents the development of a novel flexure-based micromanipulator for

the alignment of optical components (Figure 1). The realized device is based on a six-

axis parallel mechanism and piezo-positioners using stick-slip principle and providing

nanometre resolution. Results of this paper show that the flexure-based design allows

realizing nanometre steps with the moving platform in all axes.

1 Introduction

Laser technology had a large impact on industrial and everyday applications within

the last decades. The increasing demand for laser systems requires automation of

manual alignment and assembling processes to provide low cost high quality

solutions. In complement to existing robot assembling systems like SCARA

kinematics with large workspaces and low positioning accuracy, manipulators for

precise component alignment with six degrees-of-freedom are necessary.

Figure 1: Design of six-axis micromanipulator

268


For this purpose, a symmetric, hybrid serial-parallel structure has been chosen that

consists of three inextensible struts which connect three non-collinear points of its

moving platform to its base. The motion of the manipulator is obtained by moving the

lower ends of the struts on the base plane by means of three identical trays [1]. This

strutucture is analysed in terms of geometrical influences on relevant performance

characteristics, such as stiffness, singularities, workspace and transmission behaviour.

The results are used to develop a systematic, graphical optimization technique for the

parallel strucutre. For the flexure design joint angles and deformation forces are

derived. Two universal joints were integrated in the kinematic chains instead of

revolute and spherical joints presented in an earlier paper [2]. A method was

developed to miniaturize the micromanipulator for a given workspace by

optimization of flexures considering results of fatigue test for the joints.

2 Flexure hinges

The mathematical analyses resulted in required joint displacements of more than

±15°. For such large displacements a special type of flexure design has been chosen

that differs from the wide-spread and well-known notch type joints or leaf springs.

The chosen flexure type is a torsional joint with cross-shaped cross section (Figure 2)

which generates pure rotational motion with widely reduced axis drift as well as a

superior off-axis stiffness compared to many other types of flexure designs [3] [4]. It

has further been optimized to achieve the best performance in terms of precision of

motion and off-axis stiffness.

Figure 2: Influence of geometric parameters on maximum joint angle (left side);design of universal joint (right side)

Detailed results of the flexure design include the comparison of analytical

calculations with numerical simulations (stiffness and stresses) as well as

experimental fatigue tests to find an optimum design within the given constraints. As

269


a part of the micromanipulator design both optimization methods have been

integrated considering resulting space restrictions and drive specifications.

3 Workspace

The achieved overall size of the micromanipulator is Ø 115 mm × 110 mm (Figure 1)

for a coupled workspace of ±1 mm in XYZ and ±1° in ΨΘΦ (Figure 3) and a

compliance of the structure in vertical direction of 2.5 µm / N.

Figure 3: Translational workspace

4 Motion resolution and repeatabilty

The repeatability and the resulting motion resolution (Figure 4) of the compliant

mechanism have been characterised by means of interferometer measurements. The

bidirectional repeatability of the micromanipulator is around 50 nm.

Figure 4: Motion resolution measurements

5 Conclusion

The presented six-axis manipulator design was optimized for assembling various

optical laser components e.g. FAC lenses or laser resonators. Piezo-positioners and

270


flexure hinges without hystersis effects provide highly precise motion of the moving

platform in nanometre resolution.

In addition to the presented manipulator a flexure-based micromanipulator with three

degrees-of-freedom was developed (Figure 5). These degrees-of-freedom are tip, tilt

and piston motion of the moving platform. The device has a centre camera for

component detection, a clamping unit for fixation of the moving platform and the

hinges’ protection for automated gripper exchanges. Its kinematical concept was

presented by Tahmasebi [5]. Reducing the amount of actuators causes economic

advantages for applications which require only three degrees-of-freedom.

Figure 5: Modular design of three-axis micromanipulator

References:

[1] Tahmasebi, F.: Kinematic synthesis and Analysis of a Novel Class of Six-DOF

Parallel Minimanipulators. Dissertation. Thesis Report Ph.D. Institue for Systems

Research, The University of Maryland. 1992

[2] Brecher, C.; Pyschny, N.; Souza, D. F. de: Six-axis compliant manipulator for

laser assembly. In: Proc. of the euspen International Conference, San Sebastian,

Spain. June 2009

[3] Kota, S.; Monn, Y.-M.; Trease, B.: Design of Large-Displacemment Compliant

Joints. In: Journal of Mechanical Design. Vol. 127. 2005 No. 4. pp. 788 – 798

[4] Moon, Y.-M., Kota, S.: Design of compliant parallel Kinematic Machines. In:

Proc. of ASME 2002 IDETC/CIE 2002. Montreal, Quebec, Canada, 29th Sept. –

2nd Oct. 2002. pp. 35-41

[5] Tahmasebi, F.: Kinematics of a New High-Precision Three-Degree-of-Freedom

Parallel Manipulator. In: ASME Journal of Mechanical Design. 2007. No. 129. pp.

320 – 325

271

Session 4: High Precision Mechatronics

273


FEM model based POD reduction to obtain optimal sensor

locations for thermo-elastic error compensation

J. van der Sanden, P. Philips

Philips Innovation Services, The Netherlands

[email protected]

Abstract

To compensate thermo-elastic deformations in precision systems Error Compensation

Models (ECM) can be used that predict thermo-elastic deformations based on

measured temperatures. This thermal ECM is basically a matrix representing the

relation between the temperature readings of a number of sensors on a structure and

the position shift of the point(s) of interest. An accurate ECM requires determination

of the right number of temperature sensors and the selection of good temperature

sensor locations. For this, dominant temperature shapes describing the thermal

behaviour of the system can be employed. These dominant temperature shapes result

from a Proper Orthogonal Decomposition (POD) of temperature data obtained from

simulations with a FEM model using time dependent loads. A new algorithm is

proposed to select sensor locations from all nodal locations in the FEM model that

can properly identify the set of dominant temperature shapes. Furthermore, a reduced

FEM model approach is used to make POD decomposition practically possible as

well as frequency domain evaluation of the thermo-elastic ECM.

Water cooling

Linear motor coils

without magnet track

Triangular plate

Stage frame

Point of interest

X+Y+

Fig.1 Precision stage test set-up Fig.2 FEM model stage frame and motors

275



1 Introduction

The work has been applied on a precision stage that basically consists of a stage

frame and two linear motors connected to the side of the frame (see figure 1).

Furthermore, a triangular plate is mounted on the frame. In the middle of the

triangular plate a rectangular bar is connected which is pointing downwards. The end

of the bar is the point of interest. The stage frame is conditioned with cooling water

running through two channels on both sides of the frame.

2 FEM model based POD reduction

The FEM model that has been made of the stage application is limited to the stage

frame with its linear motor coil assemblies (see figure 2). The model is used to

calculate the temperatures as function of time with known motor loads. A temperature

identification matrix is generated from all nodal temperatures at each time instant.

This identification matrix can be decomposed as the product of three other matrices

using POD also known as SVD (Singular Value Decomposition). The columns of the

first matrix UT are m independent orthogonal POD temperature shapes. The second

matrix T is basically a diagonal matrix with Singular values indicating the

importance of each POD shape. Since obtaining the temperature identification matrix

for all nodes, and decomposing it using the POD algorithm is very demanding in

terms of computation time and memory, the FEM model has been reduced using

Arnoldi reduction [2]. This made it possible to solve the thermal problem with only

120 Arnoldi states X and still obtaining results which are accurate for load fluctuation

up to 1 Hz. The Arnoldi states X and FEM model temperatures T are related by the

Arnoldi projection matrix V: XT V .

Now an Arnoldi state identification matrix is calculated using the reduced model.

This matrix is decomposed resulting in the matrices: UX, X and WX. Here the shape

matrix UX is of dimensions pxp with p the number of Arnoldi states. It can be proven

that the first p POD temperature shapes of UT can be approximated by: XT UVU

~ .

From the POD temperature shapes TU

~ the most relevant shapes, with the largest

Singular values, are selected. Furthermore, the displacement sensitivity of the point of

interest for thermo-elastic deformations of each relevant POD temperature shapes is

276


investigated by enforcing the POD temperature shapes as a prescribed temperature

condition on the FEM model and calculating the deformations. The overall shape

importance can be expressed by the product of Singular value and shape sensitivity.

3 Sensor placement

Next the proper sensor locations need to be determined to identify the selected POD

temperature shapes. We want to select the sensor locations from all nodal locations of

our FEM model. As a criterion to optimize the sensor locations we use the condition

number of the matrix formed by the coefficients of the selected POD shapes at the

nodes of the selected sensor locations (see [1] for details). This criterion helps to

come up with an ECM with relatively small coefficients which limits the effect of

temperature measurement inaccuracies and sensor noise. Calculating the condition

numbers for all m possible (nodal) sensor locations is practically impossible due to

the combinatorial nature of the problem, so an efficient algorithm is necessary.

The algorithm described in [3] focuses on modal shape identification of structural

dynamics. Although the algorithm seemed to be appropriate for our problem, it did

not yield very good results in our case. In order to solve this problem we propose a

new algorithm that determines sensor locations resulting in a low condition number.

3.1 Algorithm

The temperatures of each POD temperature shape are sorted with respect to their

absolute value. The FEM nodes are sorted accordingly. This results in a matrix of

node numbers of which the columns are the sorted nodes of each POD shape. For the

nodes in each row the condition number is calculated. The nodes in the row with the

smallest condition number are selected as sensor locations.

0 0.05 0.1 0.15 0.2 0.25 0.3 0.35-0.1

-0.05

0

0.05

0.1

0.15

0.2

0.25

0.3

0.35

0.4

12

3

4

5

6

0 500 1000 1500 2000 2500 3000 3500 4000

-5

0

5x 10

-7 Error correction performance X

Time [s]

Positio

n c

ha

nge

[m

]

X

Correction

Error

0 500 1000 1500 2000 2500 3000 3500 4000-10

-5

0

x 10-7 Error correction performance Y

Time [s]

Positio

n c

ha

nge

[m

]

Y

Correction

Error

Fig.3 Selected sensor locations Fig.4 Correction model performance

277


The algorithm has been applied on a subset of nodes limited to the frame surface.

This resulted in a sensor location set (see figure 3) with a low condition number.

4 Error Correction Model

The ECM is derived by means of a least square fit applied on simulated time

dependent temperature data at the selected sensor locations and the corresponding

displacements at the point of interest. Note, that also measured data can be used.

Next, the predicted XY-position changes of the frame are calculated and compared

against the simulated XY-position changes (see figure 4). This shows small

differences of less than 8% for X-position changes and less than 1% for Y-position

changes. A more general approach to evaluate the ECM is to investigate the

performance in the frequency domain. Figure 5 shows the frequency response

function (FRF) for the motor load variation. The FRF shows that the ECM (--)

performs reasonable well for motor load variation over the whole frequency range of

interest. Similar, the performance w.r.t. cooling water temperature (see figure 6) and

ambient temperature variation can be evaluated.

10-6

10-4

10-2

100

102

0

1

2

3

4x 10

-8

frequency [Hz]

Sen

siti

vity

[m/W

]

FRF motor load variation, -simulation --correction model

10-6

10-4

10-2

100

102

-600

-400

-200

0

200

frequency [Hz]

pha

se [

deg

]

X/Qmotor

Y/Qmotor

X/Qmotor

Y/Qmotor

10-6

10-4

10-2

100

102

0

1

2

3

4x 10

-8

frequency [Hz]

Sen

siti

vity

[m/1

0m

K]

FRF water temperature variation, -simulation --correction model

X/Twater

Y/Twater

10-6

10-4

10-2

100

102

-400

-200

0

200

frequency [Hz]

pha

se [

deg

]

X/Twater

Y/Twater

Fig.5 Frequency response to motor load Fig.6 Frequency response to water temp.

References:

[1] A.H. Koevoets et. al. ‘Thermal-Elastic Compensation Models for Position

Control’, ASPE 2009

[2] I.M. Elfadel and David D. Ling, ‘A block Rational Arnoldi Algorithm for

Passive Model-Order Reduction of Multiport RLC Networks’, Proc. ICCAD, 1997

[3] C. Stephan, ‘Sensor Placement for Modal Identification’, Mechanical Systems

and Signal Processing 27, (2012) 461-470

278


2-DoF magnetic actuator for a 6-DoF stage with long-stroke

gravity compensation

R. Deng, J. W. Spronck, A. Tejada, R. H. Munnig Schmidt

PME: Mechatronic System Design, Delft University of Technology, the Netherlands

[email protected]

Abstract

High-precision positioning systems, such as lithography wafer scanners, vibration

isolators and gravity compensation systems require multi-DoF stages. These stages

generally apply magnetic actuators because of their contactless operation and high-

force capacity. The working range of current magnetic actuators in the levitating

direction is limited to around 1mm [1]. However, in some applications such as wafer

loading in nanoimprint lithography, a long-stroke motion is required. Although

increasing the airgap width would increase the working range, it would also require a

larger driving current and, thus, more heat dissipation, which is undesirable for high-

precision systems. To alleviate this problem, the design of a novel 2-DoF magnetic

actuator is presented in this paper. The actuator, presented in Section 1, is capable of

both long-stroke (20mm) and short-stroke (2mm) motions in two perpendicular axes.

In the long-stroke direction the actuator can achieve high-precision positioning with

low power and a tuneable constant force, which is confirmed both by simulation and

experiments. In the short-stroke direction, it works as a conventional reluctance

actuator. Moreover, as shown in Section 2, the actuator could also be used to design

6-DoF maglev positioning stages with gravity compensation (see Figure 1).

1 Basic configuration and working principles of the 2-DoF actuator

The actuator consists of an iron mover and an iron C-core stator with two permanent

magnets and two coils (Coil1 and Coil2), as shown in Figure 2. The two permanent

magnets have the same orientation in the X-axis and provide a static force that allows

for gravity compensation with minimal power consumption (i.e., low coil current),

thus reducing the heat in the system. Coil2 provides a 2mm short-stroke conventional

reluctance actuation in the X-axis with initial negative stiffness, while Coil1 provides

a 20mm long-stroke actuation in the Z-axis (dynamic force) with initial low stiffness

over the full stroke.

279



A1 A3A2

z

x

Y

Mask

Wafer

6-DOF stage

Figure 1 Figure 2

Iron mover

Coil1

Coil2

Iron stator

Permanent

magnets

z

x

Figure 1: Proposed 6-DoF high-precision positioning stage concept using three 2-DoF

actuators. The stage is capable of a 20mm stroke in the Z-axis with gravity compensation and a

1mm stroke in the XY plane.

Figure 2: Proposed 2-DoF actuator configuration with iron stator and mover. The two

permanent magnets on the stator are in the same orientation allowing for gravity compensation.

Coil1 and Coil2 are both mounted on the stator allowing for active mover position control along the Z and X axes, respectively.

1.1 Gravity compensation and actuation in Z-axis

The total flux used for Z-axis levitation and actuation is the sum of the fluxes of the

two permanent magnets and that of Coil1. The latter changes the field strength in the

airgap to achieve different levels of constant force or to actuate in the Z-axis. On the

assumption that there is no stray flux, no saturation, nor iron reluctance, the levitation

force in the Z-axis can be derived from the magnetic energy stored in the airgaps as

follows: The effective airgap volume ( )g d gV y z x , wheredy z is the overlap surface

between the magnet and the mover (the actuator thickness out of plane × the overlap

length between the magnet and the mover), gx is the total airgap.

The total magnetic energy stored in the airgap is 2

0

g

m g g g d

BE B HV x y z .

Here, 0is the magnetic permeability in vacuum, H is the magnetic field strength,

and gB is the flux density in the airgap given by

0 2

r m r mg

t g m g

B l B lB

R A l x

,

where , rB ,

gA , and tR are, respectively, the loss factor, the remnant flux density,

the overlap magnet surface with the mover, and the total reluctance. The latter is

given by

0 0 0

22

g m gmt m g

d d d

x l xlR R R

y z y z y z.

Thus, the magnetic force in the Z-axis is 2 2 2 2

0 0 (2 )

gm r mz g d g d

m g

BE B lF x y x y

z l x.

280


Figure 3: Fz COMSOL simulation with no coil

activated. From Z-position 2mm to 18mm the force stays around 8N (Y-axis in COMSOL

represents Z-axis here).

Figure 4: Fz in experiments. The middle line is the force when no coil is activated. The ones

above and below are forces with current in

Coil1 in two directions.

The force Fz as function of the Z-position is modelled with real material parameters

by means of 2D COMSOL and the results in Figure 3 show a nice flat top at 8N. This

implies an initial low stiffness and low-power gravity compensation for 0.8kg. Figure

4 shows the experimental validation using a linear stage and a force sensor: Fz was

measured at 20 Z-positions and 3 Coil1 current levels (0A, ±0.3A), showing that Fz

can be both increased and decreased by varying the current in Coil1. The force level

(gravity compensation) and stiffness in the Z-axis can also be tuned by modifying the

stator or/and the mover geometries around the airgap. The Z-force profile also can be

shaped by locally modifying the permanent magnets field strength. Finally, the

efficiency of the X-axis actuation could be increased with additional iron paths.

1.2 Actuation in X-axis

The 2-DoF actuator works as a conventional reluctance actuator in the X-axis. It has

negative stiffness and a lateral force, Fx, which is quadratic with the current and the

position. As Figure 5 shows, Fx presents a large linear range around the working

point (the middle position) because of the actuator symmetry. The total airgap in the

X-axis is 2mm. Coil2 can actively control the mover position in a range of ±0.5mm

around the central working point.

2 6-DoF positioning stage concept

Figure 1 shows a first concept for a 6-DoF stage using three described 2-DoF

actuators (A1, A2, A3). This stage would allow a 20mm stroke in the Z-axis of which

about 10mm with gravity compensation, initial low stiffness in Z, Rx and Ry. In the

XY plane the stage would have a 1mm stroke with initial negative stiffness.

281


Figure 5: Fx COMSOL simulation. At 10mm Z-position, with 100 windings, 0A, ±1A, ±2A

current in Coil2, the mover can work in the range of ±0.5mm around the working point.

The stage position could be controlled using a 6-DoF MIMO controller and a 6-DoF

laser interferometer measuring system. In case of a limited planar working range,

three optical encoders measuring the long-stroke displacement and three capacitive or

inductive sensors measuring short-stroke displacement could be an alternative.

The 6-DoF stage could be used for nanoimprint, using Z-axis actuation to generate

both the printing and releasing forces. In such case, accelerometers would be needed

to control the releasing force which is known to be impulsive [2]. Additionally, the

current in the coils could be used to infer the forces in the Z-axis, which could be

used in both feed-forward and feedback control.

3 Discussion and conclusion

The proposed 2-DoF actuator can achieve tuneable constant force in the Z-axis with a

long stroke. The force level in the whole working range can be tuned either by

changing the current in Coil1 or by modifying the stator or/and the mover geometries

around the airgap.

The demonstrated 2-DoF actuator is easy to design and build and can be used as a

flexible component for a 6-DoF positioning system. Another application could be a

vibration isolation stage because of its initial low stiffness in the Z-axis. This and

other applications and the tradeoffs between actuator and control design are currently

under investigation.

References:

[1] A.T.A. Peijnenburg, J.P.M. Vermeulen, J. van Eijk. Magnetic levitation systems compared

to conventional bearing systems, Microelectronic Engineering 83 (2006) 1372-1375

[2] H. Atasoy, M. Vogler, T. Haatainen, et al. Novel thermoplastic polymers with improved

release properties for thermal NIL, Microelectronic Engineering 88 (2011) 1902–1905

282


Highly accurate passive actuation system

S. A. J. Hol 1, J. Huang1, W. Zhou1, M. Koot1, H. Vermeulen1, J. van Eijk2,

R. Munnig-Schmidt3 1ASML BV, Netherlands 2MICE BV, Netherlands 3Delft University of Technology, Mechatronic Systems Design, Netherlands

[email protected]

Abstract

This paper describes an alternative method to drive and control the fine positioning

(short stroke) stage of a precision positioning device. The short stroke actuators are

replaced by a pair of passive magnetic springs. Now the control forces are determined

by the relative displacements between the long-stroke (coarse positioning) and short-

stroke modules, while it eliminates large actuators and power dissipation in the short-

stroke. This approach requires a special control effort for the long-stroke actuation

system.

Introduction

Precision positioning stages commonly consist of a combination of a coarse

positioning module with limited accuracy (long-stroke), at which a fine positioning

module (short-stroke) is cascaded. Figure 1 illustrates this principle. The precision of

the system is achieved with a fast linear short-range actuator, while the coarse

positioning module keeps the short-range actuator in its optimal working range.

Accurate positioning

relies on accurate control

and effective reduction of

disturbances. Both

positioning modules use

electromagnetic actuating systems. Especially the actuator for the fine positioning

module must generate high forces at high accuracy, which is a contradicting

requirement.

corresponding author

Figure 1 Cascaded modules

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Figure 2 Spring element between cascaded modules

Over the past decades a steady increase in required accuracy and speed is observed

with a proportional increase of accelerations. This requires the actuators to be large

and heavy while a substantial amount of heat is dissipated, significantly affecting the

thermal stability of the application and jeopardizing the system’s positioning

accuracy, since it leads to unwanted structural deformations.

1 Passive Drive System - Theory

This paper investigates

the possibility to make

use of the actuators of

the coarse positioning

module to accelerate

the fine positioning

module as well, thereby reducing the heat load of fine positioning actuators. The

force is transferred by an inserted spring element between the coarse and fine

positioning modules as shown in Figure 2 as proposed in [1]. Main advantage of a

spring is that most of the force is delivered by the spring. Now only a small,

lightweight actuator is needed for counteracting unknown disturbance forces.

1.1 Contact less non-linear spring

The most straight-forward implementation would be an ordinary mechanical spring,

but that would lead to too large

disturbing forces between the coarse

and fine positioning module.

Therefore a magnetic, nonlinear,

contact-less spring is proposed for

this development as shown in Figure 3. This spring consists of two pairs of repelling

permanent magnet arrays.

Figure 3 Magnetic arrays

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The benefit of such a non-linear magnetic spring element is the minimum stiffness

behavior in the mid position of the spring as can be concluded from the force-stroke-

characteristics shown in Figure 4. At its neutral position (mover displacement equals

0 mm) the force is zero and also the slope of the graph (which represents the

stiffness) reaches its minimum value. This configuration effectively reduces the

disturbance forces transmitted by the coarse positioning module around the neutral

position.

1.2 Control strategy

Driving this system requires a renewed control strategy [2]. With a regular

electromagnetic actuator the force on the fine positioning module is linear

proportional to the current through the actuator. For this application, the force is

depending on the relative position of the fine positioning module with respect to the

coarse positioning module. The fine positioning module has to track a certain set-

point and the corresponding trajectory for the coarse module can be computed using

an inverse model of the (non-linear) spring. For a standard set-point, the required

trajectory for the coarse positioning module will exceed the constraints on actuator

forces. A strategy is developed to derive a set point that minimizes the time from

Figure 4 Force-stroke-characteristics, both measured and simulated

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initial to end state, satisfying dynamic model equations and respecting the limitations

on force, velocity, acceleration, jerk and snap of the coarse positioning module [4].

2 Passive Drive System - Practice

Figure 5 shows the hardware for the prototype of the design [3]. The fine positioning

module is suspended by a 5-DOF air bearing system. The position measurement

between fine and coarse positioning module is accomplished by an optical encoder.

The coarse positioning module was mounted on two powerful linear actuators (not

shown in the picture).

The system has been subjected to a repetitive motion profile with a peak velocity of 1

m/s and peak accelerations of 100 m/s2. The tracking error of the fine positioning

module during constant velocity was less than 2 micrometer as shown in Figure 6.

This is equal to the noise level of the used sensors to measure the position. Therefore

the measurement system is currently being improved. This system shows high

potential to even reduce its tracking error significantly to the nanometer range.

Figure 5 Hardware

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3 Conclusions

In this project has proven that transfer of driving forces from a coarse motion system

to a fine positioning module can be done with passive elements. This has been proven

by theoretical modelling and measurements on a prototype implementation.

Additionally, the non-linear spring, created by opposing permanent magnets, is well

suited for this purpose. Finally, suitable trajectories for the motion elements can be

derived and excellent motion performance can be obtained.

References:

[1] Passive Stage Actuation, S.A.J. Hol et al., internal ASML

[2] Control strategy for Passive Short Stroke Driving, M. Koot et al., internal ASML

[3] Mechanical Design of PSS, J. Huang et al., internal ASML

[4] Passive Stage setpoints, J. van Eijk, internal ASML

Figure 6 Systems response to motion profile

287


Design and fabrication of a novel centimeter scale three

dimensional silicon tip, tilt and piston mirror mechanism

J. Kruis1,2, F. Barrot1, L. Giriens1, D. Bayat1, R. Fournier1, S. Henein2, S. Jeanneret1 1Centre Suisse d’Electronique et de Microtechnique (CSEM), Switzerland 2École Polytechnique Fédérale de Lausanne (EPFL), Switzerland

[email protected]

Abstract

A novel centimetre scale tip tilt piston mirror mechanism has been designed in

silicon. The mirror consists of three identical mechanism parts and one mirror part.

The fabrication of the parts was done with photolithography and Deep Reactive Ion

Etching and they are presently being assembled. The originality of the proposed

concept resides in breaking down of the kinematic structure into three identical planar

flexure-based monolithic structures and the isostatic alignment concepts used to

assemble these planar structures into a three dimensional structure.

1 Introduction

The up scaling of silicon Micro Electro Mechanical Systems (MEMS) to millimetre

and centimetre scale is a bottom-up approach of coping with the challenges between

the domains of MEMS and classical metal-based precision mechanisms. Basing this

approach on silicon presents several advantages such as the absence of fatigue,

machining accuracy (typically one order of magnitude better than that of Wire-EDM),

possible integration of sensors and actuators inside the articulated structures

themselves and batch production on wafers.

However, the fact that most mature Silicon processing is planar (2D) or stacked

planar (2.5D) in combination with silicon being a brittle material, prevents silicon

from being widely used at the centimetre scale. The challenge of assembly has been

coped with in various publications [1], [2] but typically on a micro scale. We also

coped with these challenges in our previous research [3].

We characterised the effects of stress concentration, crystalline orientation and

surface treatments on the fracture strength of silicon flexures. This led to a set of

288



fabrication and design rules enabling the design and fabrication of more robust parts

which are less prone to fracture.

In the same research, we also showed that, assembling centimetre scale silicon slabs

in three dimensions is feasible, and can lead to the production of silicon based three

dimensional mechanical systems. In this context, we designed, assembled and

characterized a silicon sugar cube size delta robot as displayed in figure 1. This robot

consists of three identical flexure based silicon slabs assembled together to form the

architecture of a delta robot.

Figure 1: Photos of a flexure slab of the sugar cube Delta-Robot (left) and the

sugar cube size Delta-Robot (right) [3]

2 The tip tilt piston mirror mechanism

In an effort to further improve our expertise in the assembly of 3D centimeter scale

silicon structures comprising delicate flexure mechanisms, a novel silicon Tip, Tilt

and Piston mirror mechanism (TTPmm) has been designed and produced. The

originality of the proposed concept resides in the breaking down of the kinematic

structure into three identical planar flexure-based monolithic structures and the

isostatic alignment concepts used to assemble these planar structures into a three

dimensional structure.

The achieved device has a relatively small volume compared to classical metal-based

precision mechanisms while achieving large displacement ranges compared to

classical MEMS.

2.1 Fabrication

The 2.5D silicon parts were micro-fabricated using photolithography and deep

reactive ion etching (DRIE). For the fabrication, a Silicon On Insulator (SOI) wafer

(500 μm handle layer, 2 μm buried Oxide, 45 μm device layer) was used. In addition,

for the mirror surface and solder pads, a 200 nm gold layer was deposited.

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2.2 Kinematics

The TTPmm consists of a total of four monolithic silicon components: three identical

silicon slabs containing the kinematic structure (hereafter referred to as “the flexure

slabs”) and the silicon mirror. The flexure slabs have multiple functionalities:

decoupling the actuators; providing a translation of the integrated mirror frame.

The kinematics of one flexure slab is displayed in figure 2. The structure consists of a

linear guide which is actuated through a silicon rod used to decouple parasitic

motions of the actuator. The linear guide pushes the mirror frame with a rod in the Z

direction while another rod constrains the mirror frame in X.

Figure 2: Kinematics of a flexure slab.

The implemented design of the TTPmm is capable of +/- 4 Deg (Tip and Tilt)

rotations in the mirror plane; the (Piston) translation range out of the mirror plane is

+/- 0.6 mm.

2.3 Proposed Assembly strategy

For the assembly, each of the three flexure slabs constrains the mirror in 2 Degrees of

Freedom (DOFs): 1 axial and 1 azimuthal DOF. With the introduced method for

alignment we achieve isostatic positioning of the mirror. The flexure slabs are fixated

to a metal frame.

The entire assembly fits in a 40x40x42 mm3 rectangular volume and is shown in

figure 3. The actuators are in turn coupled to the silicon rods to provide the actuation.

The flexure slabs are fixated with either gluing or soldering. To allow for soldering

the flexure slabs and the mirror have a gold coating on their interfaces.

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Figure 3: Tip Tilt Piston Mirror Mechanism

2.4 Actuation and sensing

Three commercial ultrasonic piezo actuators were selected to provide the forces for

actuating the TTPmm. However, the design allows for various linear actuators.

Although sensing functions are not directly implemented in the current design, a first

approach is to use an optical measurement system based on the mirror itself to

characterize its displacements.

3 Conclusion

The design demonstrates the possibilities of creating centimetre scale 3D assembled

silicon mechanisms by use of parts integrating novel features both for alignment and

for fixation. Presently the parts have been fabricated and we are in the process of

assembling and characterizing the demonstrators.

Typical application areas of such an centimeter scale opto electro mechanical systems

include laser machining, scanning Light Detection And Ranging (scanning LIDAR)

systems, as well as pick off mirrors for multi object spectrometers.

References:

[1] High Yield Automated MEMS asssembly, D.O. Popa et al., Conference

proceeding, CASE September 2007

[2] Robotic Microassembly of 3D MEMS Structures, N. Dechev et al, Book, chapter

6, pp. 225 – 248, 2009, ISBN 9780470484173

[3] Silicon flexures for the sugar-cube delta robot, S. Henein et al., Conference

proceeding, EUSPEN May 2011

291


Superstructures control with active tie rods

C. Collette, D. Tshilumba, L. Fueyo-Rosa

University of Brussels, Belgium

[email protected]

Abstract

In this paper, we explore the possibility to increase the stability of particle collider

superstructures with a network of active tie rods. Basically, they consist of carbon

fibre tie rods, fixed at one end on the superstructure, and at the other end to stable

points (e.g. detector frame) through active tendons. In the first part of the paper, the

solution has been tested on a finite element model of one half of the future Compact

Linear Collider (CLIC) final focus structure. With a reasonable design using four

rods, it is shown numerically that the compliance is decreased by at least a factor 4,

i.e. that the structure is 4 times more robust to technical noise at low frequency. It is

also shown that the active rods offer two additional important advantages. The first

one is that they can be used to damp significantly all the modes observable by the

tendons. The second one is that they can be used to realign the superstructure

components. The second part of the paper presents a successful experimental

validation of this concept, applied to a scaled test bench. The bench has been

designed to contain the same modal characteristics as the full scale superstructure. It

is shown that the superstructure compliance can be decreased by a factor 30 in a large

frequency range, and locally by nearly three orders of magnitude. The capability of

the active tendons to damp and move the structure is also demonstrated

experimentally, and found to comply well with theoretical predictions.

1 Introduction

Sometimes, in several large experimental facilities, precise equipments have to be

mounted on very large structures. Unfortunately, these structures do not represent a

very stable support, and can significantly affect the stability of the equipments, and

thus, also affect the quality of the experiments. An emblematic example is the so-

called final focus of future linear particle collider, where the electromagnets

dedicated to focus the beams of particles are supported by large cantilevered

292



structures, called the superstructures in this paper. A representative example of such

structure exist at CERN (CMS experiment), and is shown in Figure 1. Recent

measurements on this structure [1] have shown large vibrations

Figure 1 Superstructure of the CMS experiment at CERN.

of the free end of the superstructure (about 90 nm, i.e. nearly 3 orders of magnitude

above the stability requirements), which are also poorly correlated with ground

motion. This indicates that the structure is too sensitive to direct disturbing forces

(ventilation fans, cooling, electronics, acoustic noise..). In order to address this issue,

we propose to reinforce the structure with a network of carbon tie rods, as presented

in the next section.

2 Superstructure stabilization strategy

The model of the CLIC final focus superstructure is shown in Fig.2. It consists of a

large tube, cantilevered on the tunnel wall, inside which the electromagnet is

supported. The compliance of the free end of this structure is shown in Fig.2. Using a

finite element model of this structure, we have calculated that a network of 4 tie rods

(shown in red in Fig.2), with a diameter of 4 cm, can reduce the compliance by a

factor 4 in directions perpendicular to the tube main axis [2].

Now, let us further consider that an active tendon (constituted by a force sensor in

series with a displacement actuator) is fixed at one end of each tie rod, and that we

use decentralized loop in each tendon, with the following controller:

H=gs/(s+a)2,where s is the Laplace variable, a is a parameter and g is the gain.

It can be easily shown that the stiffness matrix K of the system becomes

K+kcLLT(s+a)

2/[(s+a)

2+gs]), where kc is the rod stiffness and L is a

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matrix projecting the rod forces in the structural degrees of freedom. At low

frequency, the matrix is unchanged by the controller, i.e. it does not affect the

robustness to the disturbing forces. At high frequency, the second term is proportional

to s, which creates an active damping (also visible in Fig. 2). Obviously, the active

tendons can also be used to realign the superstructures with a good authority.

Figure 2: 3D view of the final focus superstructure and its compliance at the free end

in the vertical direction.

3 Experimental validation

Figure 3(a) shows a picture of the experimental set-up and Fig. 3(b) shows a zoom on

an active tendon. The detector has been represented by a rigid frame. The compliance

has been measured in the vertical and lateral direction by exciting the structure with

an instrumented impact hammer and measuring the displacements at the same

locations. It is shown in Fig. 3(c) for the vertical direction. One sees that the

compliance has been divided by a factor 30 at low frequency and, around 30 Hz, by

more than two orders of magnitude. Theoretically, a higher reduction of the

compliance could be obtained by an additional tension in the cables. However, an

excessively high a value of the tension becomes risky for the force sensors.

Finally, the capability of the actuators to move the free end of the tube has been also

successfully verified by injecting out of phase sinusoidal signals in the two vertical

actuators (not shown in this paper because of space limitation).

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Figure 3. (a) Scaled flexible structure with active cables; (b) Active tendon; (c) Effect

of the active cable on the compliance of the free end of the tube.

4 Conclusions

In this paper, it has been proposed to reinforce the superstructure with a network of

active tie rods. Using a realistic design, it has been shown numerically that the

compliance of the superstructure can be reduced by a factor 4, with only four tie rods.

In addition to stiffening, it has been shown that the structural damping can be

significantly increased with an active tendon connected at one end of each tie rod. A

third property of the active rods network is that it can also be used to realign the

superstructure. These results have been confirmed experimentally on a scaled test

bench. It has been demonstrated that a network of four cables decreases the

compliance of the test bench by a factor 30. The capability of the active tendons to

increase the structural damping and to reposition the structure is also confirmed

experimentally, and found to comply with the theoretical predictions.

References:

[1] Collette C., Janssens S. and Tshilumba D., Control strategies for the final focus

of future linear particle collider, Nuclear instruments and methods in physics

research section A , vol.684, 7-17 (2012).

[2] Collette C., Tshilumba D., Fueyo-Rosa L. and Romanescu I., Conceptual design

and scaled experimental validation of an actively damped carbon tie rods support

system for the stabilization of future particle collider superstructures, Review of

Scientific Instruments, vol.84(2), 023302 (2013).

295


Modelling lateral web dynamics for R2R equipment design

B. J. de Kruif and H. E. Schouten

TNO, Netherlands

[email protected]

Abstract

A model describing the lateral dynamic behaviour of a web between and on rollers is

implemented to aid in the design of roll to roll equipment. The model is validated on

an industrial setup, and disturbances acting on the web were estimated. The model

equations were used to visualise the interdependence between the span length and the

disturbances allowed. This resulted in a span length that was least sensitive to

disturbances.

1 Introduction

Roll to roll production of solar cells, displays, and printed electronics is a likely

approach to meet the high throughput demand in the future [1]. A plastic or metal

web passes several processing stations, e.g. printing, lamination and slitting, to come

to an end product. The current position accuracy of these processes is typical worse

than 25 μm, while the alignment of the different processing steps has to be an order of

magnitude better than what the state of the art equipment offers, to meet the future

needs [2]. High accuracy web handling, as well as a correct mechatronic design, are

needed to meet the allowed lateral and longitudinal displacement specifications and

to minimise the internal stresses in the web. Misaligned rollers will introduce tension

differences along the width of the web that can become larger than the longitudinal

tension. As a result, the web will be in compression at a side, and a ‘bag’ will occur.

In this work we investigate how accurate two adjacent rollers need to be aligned and

how large disturbances can be before the web starts to show bags or other unwanted

behaviour shows. This is investigated by implementing a model of the web

behaviour. With this model a trade-off between span length, roller alignment and

disturbances is made.

2 Modelling

The shape of the web between two roller as function of time is calculated in two

steps: first, the motion equations describing the position of a web at the exiting roller

296



are derived [3]. These equations describe the lateral acceleration as function of the

entering angle and position, and the motion of the rollers. With this equation the web

angle at the exiting roll can be calculated too. Secondly, with the web position and

angle at the entering and existing rollers known, the shape of the web is determined

between the rollers. The shape changes instantaneously by changing the web position

or angles at the rollers, which assumes the web to be massless. The deflection of the

web’s centre-line is described by

, in which is the deflection, the

along-axis and a parameter depending on the material properties, dimension and

tension.

As introduced previously, no part of the web should be in compression. When the

web is entering the span at an angle, whether this is due to a steering action or due to

misaligned rollers, moments are introduced into the web. The moment results in a

non-uniform tension distribution, and when the moment is too large, one side will be

in compression. This is comparable to the behaviour of a beam. The angle at which

the web gets into compression is called the critical angle. Furthermore, an angle at the

entering web will also introduce lateral forces at the roller and a displacement of the

web at the exiting roller. These effects should be limited to avoid lateral slip and drift

respectively.

2.1 Model validation

The dominant behaviour of the model is validated on an industrial setup. A schematic

representation of the setup is shown on the left of figure 1. The edge sensors are

denoted by s1 till s5. The measured edge location for sensor s2 till s5 are given as blue

lines in the right of the figure. The displacement guide changed its angle around the

double arrow several times to excite the system.

A dynamic model is made that implements the above described approach for this

setup in Matlab/Simulink. The edge of the web was predicted at the sensors s2 till s5.

The position of the web at s1 as well as the angle of the displacement guide were used

as input for this prediction. The grey line shows the predicted lateral position at the

sensors.

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Figure 1: schematic view of the validation setup (left), and the simulation (grey) and

measured (black) web positions.

The model predicts the position of the web at the sensors well. However, the model

shows some overshoot when the displacement guide makes a large step. The angles

between the web and the rollers at these steps is larger than the critical angle,

resulting in compression. This is shown as marks at the bottom of the figure. Based

on the sensor data, the angle of the entering web is estimated with a maximum

amplitude of 1 mrad.

3 Design example

The dynamic model can be used for, e.g., concept comparison or sensitivity analysis.

In this work we use it to investigate whether a web between two rollers will come

into compression, introduces too large lateral forces on the roll, or drifts too much to

the side, i.e. if it moves stably. These three criterions can all be related to the web

angle at the entering web, and are all a function of the length between the rollers. The

calculations were done for PET foil with a width of 300 mm, a thickness of 125 μm

and a tension of 100 N. The maximal allowed lateral force was set to 10 N, with a

maximal deflection at the exiting roller of 1 mm. In figure 2 the angles are plotted

that would break one of these criterions. When the entering angle in combination with

a specific length stays within the white area, the web will behave correctly. The dark

grey area shows the angles that will result in compression, too large shear, or too

much deflection for the static situation. It shows how well the rollers must be aligned

to operate, even in the absence of disturbances. Until a certain length, increasing the

length allows for larger misalignment errors.

The web when entering a span will have angle disturbances due to upstream steering

actions. For the web to behave well with these dynamic disturbances, the magnitude

298


of the angle, has to avoid the light grey area in figure 2 too. The lines shows which

of the criterion is broken. This guides our design, as it shows what effect is

responsible for the unwanted web behaviour. In this case, an optimal length of 1.3 m

would result in the largest robustness to disturbances.

Figure 2: Allowed angles for the entering web (left), and the calculated moments

when the length/disturbance is chosen in the light grey area.

The right-hand side of figure 2 shows an example of the simulated moment at the

entering and exiting roller when the position and angle at the entering roller are

approximately equal in size as the disturbances measured at our setup. The length of

the span was chosen to be in the light grey area of the left hand figure. The figure

illustrates that the web would sometimes be in compression due to these disturbances.

This occurs if the moments exceed the middle area in the right hand side picture. The

rollers were perfectly aligned in this simulation.

4 Summary

Based on the model calculations, the interdependence between length, disturbances

and alignment errors is investigated. This showed that an optimal length could be

chosen that allowed for maximal disturbances. With the dynamic model any design

can be assessed to predict its performance and find sensitive design considerations.

Furthermore, the models can be used to develop model based controllers.

References:

[1] Schwartz, E., “Roll to Roll Processing for Flexible Electronics,” MSE 542:

Flexible Electronics, Cornell University, 2006

[2] Clemens, W. (ed), ”OE-A Roadmap for Organic and Printed Electronics, fourth

version”, Frankfurt am Main, 2011

[3] Shelton, J.J., “Lateral Dynamics of a Moving Web,” Ph.D., Oklahoma State

University, 1968.

299


Design of an active magnetic stabilizer of the dynamic

behaviour of high speed rotors

E. Brusa

Dept. Mechanical and Aerospace Engineering, Politecnico di Torino, Italy

[email protected]

Abstract

A new concept of active dynamic stabilizer to operate the rotor at fairly high

supercritical speed is here investigated. Instead of using the non rotating damping, as

the classic literature suggest, a contra-rotating action is applied. Experiments

performed on a flexible rotor being successfully stabilized by an eddy current

magnetic damper are described. Advantages of contra-rotating the control forces of

the active magnetic bearings of an electromechanical spindle are then discussed.

1 Introduction

High speed rotors are usually operated in the supercritical regime to assure a good

self-centring condition [1]. The unbalance response and the reactions of bearings are

reduced. Unfortunately, damping associated to the rotating parts of the rotor induces

some dynamic instability, above a value of the angular velocity, being referred to as

threshold of instability. Amplitude of the whirling motion exponentially grows up and

may cause some dangerous failures to the rotor. To increase the threshold, designer

applies to the stator a suitable amount of non rotating damping, being always

stabilizing. Nevertheless, if the rotor is hung on active magnetic bearings (AMBs),

the control current required to stabilize the whirling motion may be fairly large [2].

Sometimes there is no stator available to provide the non rotating damping. To

overcome those limits an innovative use of magnetic damping was investigated. A

rotating magnetic force is applied to the rotor by either an active or a passive device.

This action provides a contra-rotating damping in the reference frame of the stator,

since the rotor and the magnetic force rotate in two opposite directions. The dynamic

stability threshold of the rotor can be increased up to a fairly high supercritical regime

[3]. The so-called Jeffcott’s rotor model can be used to introduce this concept [4]. If

300



the complex coordinate z = x+iy describes the rotor radial position, the equation of

motion for a constant spin speed is:

tirnr emzcikzcczm 2 (1)

where m is the rotor mass, cr and cn, the rotating and non rotating damping

coefficients, k the stiffness and the rotor eccentricity. Rotating damping introduces a

term proportional to the radial displacement z and to the angular speed . Rotor runs

above the critical speed, cr, to reach a good self-centring and strictly below the

instability threshold, th:

m

kcr ;

r

ncrth

c

c1 (2)

Decreasing the rotating damping is often difficult, therefore a suitable non rotating

contribution is usually provided. A contra-rotating action, cd, can be used to contrast

the rotating damping, cr. It is applied in a reference frame rotating with angular speed

d with respect to the stator, so as the equation of motion becomes:

tiddrdnr emzcicikzccczm 2 (3)

If d= – and cr = cd, the instability threshold tends to infinity and all the forward

and backward whirls are stable [5]. This condition can be found evenly if dcd = –

cr. Therefore the amount of stabilizing damping can be small if a suitable contra-

rotational speed is set up. To apply this approach a contactless electromechanical

coupling is used. A first possibility is resorting to the eddy currents induced by the

rotor on a secondary contra-rotating and conductive disc or suitably modulating the

current of an active magnetic bearing, to create a contra-rotating magnetic field.

2 Eddy current contra-rotating damper

The above described approach was tested on the prototype of flexible rotor depicted

in Fig.1. A rigid frame holds up a pendulum rotor, being a disc hung to a quill steel

shaft, while at the bottom a second disc rotor is located. Rotors are separately fed by

two brushless motors. Optical sensors measure the radial displacements of the

pendulum rotor. Non rotating damping is provided by the dissipation occurring in the

supports, threaded joints and plates of the stator, while rotating contribution is

associated to the rotor clamps, shaft and disc. Damping provided by the lower disc

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looks as non rotating, if it is fixed, either co-rotating or contra-rotating if it rotates

along with the rotor or in the opposite direction. The upper disc is equipped with a

permanent magnet, which induces some eddy currents in the conductive material of

the lower disc.

d

[rpm]

d [rpm]

A B C

Figure 1: Test rig with the contra-rotating damper and experimental map of stability.

For a given gap between the two discs and when the two cylindrical whirling motions

are analyzed, the first forward mode appears stable up to 1500 rpm, when the lower

disc does not rotate. If the main rotor is kept rotating at 1500 rpm and a slow contra-

rotation of the disc is applied and increased up to 230 rpm, the forward whirl is

stabilized, but after a few seconds the backward mode is made unstable, against the

statement of the literature that backward whirls are naturally stable [1]. An

experimental map of stability was drawn (Fig.1). It shows that a contra-rotation faster

than 125 rpm causes the instability of the backward whirl (A) while a co-rotation

above 125 rpm makes unstable the forward whirl (C). Both are stable in (B).

Therefore it was found that contra-rotation increases the dynamic stability of the

rotor, provided that the instability of the backward whirl is prevented.

3 Contra-rotating damping in controlled rotors on magnetic suspension

The equation of motion of a rigid rotor suspended on AMBs, can be written as [6]:

rddrudrn FqCCKKqGCCCqM2ii (5)

being q the four translational degrees of freedom monitored by the sensors, while the

subscripts indicate non rotating, n, rotating, r, and contra-rotating, d. When the rotor

is uniquely suspended on the AMBs and a PID control is applied, actions of bearings

are non rotating and Cr and Cd vanish. The rotor is stable in the closed loop. If the

control current is made contra-rotating, Cn is converted into Cd. System is stable in

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absence of Cr. If the internal dissipation is taken into account as Cr= 70 Ns/m, the

first forward whirl becomes unstable above 15200 rpm. This threshold can be

increased if the control force is contra-rotated at 0.2 but the backward whirl

becomes unstable above 19000 rpm, being the limit of validity of the assumption of

rigid body motion. In case of the severe rotating damping condition corresponding to

Cr= 700 Ns/m, the instability threshold decreases down to 3100 rpm. A contra-

rotation at with the 70% of the action provided by the AMBs when damping was

non rotating allows having a higher threshold, at 4780 rpm, for both the forward and

the backward whirling motions, respectively. A contra-rotation at 0.5 with only

one-half the non rotating damping provided at the beginning increases the threshold

of the forward whirl to 3500 rpm, while the backward whirl remains stable. A larger

benefit in reducing the current fed to the AMBs can be found if they are used just as

stabilizing actuators on the rotor suspended on mechanical bearings.

Sensor AMB Sensor

PID control

AMB (axial)

AMB

Figure 2: Test rig of a rigid rotor upon active magnetic bearings.

References:

[1] G. Genta, Dynamics of Rotating Systems, Springer Verlag, New York, 2005.

[2] G. Schweitzer, E.H. Maslen, Magnetic Bearings: Theory, Design and

Application to Rotating Machinery, Springer, New York, 2010.

[3] E. Brusa, Stabilizer Device for Rotary Members, PCT WO/2007/122189.

[4] G. Genta, E. Brusa, Int. J. Rotating Machinery, 6(6), 2000.

[5] E. Brusa, G. Zolfini, J. Sound and Vibration, 281(3-5), pp. 815-834, 2005.

[6] N. Amati, E. Brusa, “Vibration Condition Monitoring of rotors on AMB fed by

Induction Motors”, Proc. IEEE/ASME AIM′01,Como, 8-12 July 2001, pp.750-756.

[7] C. Delprete, S. Carabelli, G. Genta, “Design, construction and testing of a five

active axes magnetic bearing system”, Proc. 2nd ISMST, Seattle, USA, 1993.

303


Physical and phenomenological simulation models for the

thermal compensation of rotary axes of machine tools

M. Gebhardt, S. Capparelli, M. Ess, W. Knapp, K. Wegener

Institute of Machine Tools and Manufacturing (IWF), ETH Zurich, Switzerland

[email protected]

Abstract

Up to now, research of the thermo-mechanical deformations was focused on the

environment, the spindle, the bed and the linear axes of machine tools. The thermal

behavior of rotary and swiveling axes was not studied in the same detail, but they are

getting more important due to the increasing requirements for 5-axis machine tools.

This paper deals with the comparison of a physical and a phenomenological

simulation model for a model-based compensation of thermal errors of rotary axes.

1 Introduction

Thermo-mechanical deformations caused by internal or external heat sources are still

responsible for up to 75% of all geometric errors on machine tools. Because of this

significance, there are many approaches with the goal to reduce these errors [1].

Regarding the thermo-mechanical flow (Figure 1), this can be achieved in two ways.

Figure 1: Thermo-mechanical flow diagram

On one hand, the causes can be minimized. This can be carried-out by reducing

power losses or by altering heat transfer in the machine tool. The temperature

distribution can also be homogenized or the resulting mechanical deformation can be

304



decreased by a more thermo-symmetrical design or a smart material mix. On the

other hand, the effect (the thermo-mechanically caused TCP error) can be reduced by

a compensation strategy. Therefore, a kind of error modeling is necessary, which

implicates proper know-how about the thermal characteristics of the relevant axis.

2 Thermal characterization of rotary axes

As mentioned in [2] and [3], an ideal measuring device for the thermal

characterization of rotary axes is the R-Test device (Figure 2, left side). When it is

carried out as “R-Test discrete” (Figure 2, right side), all significant errors of the

rotary axis or of functional surfaces can be evaluated by measuring 5 discrete points

at 0, 90, 180, 270 and 360°.

Figure 2: “R-Test discrete” setup and measuring cycle [2]

Figure 3: Axial growth of machine table (Z0T) for different rotational speeds, nmax =

maximum rotational speed.

As an example for a significant location error caused by a rotary axis, Figure 3 shows

the axial table growth (Z0T) of a C-axis machine table over 8 h (warm-up 4h, cool

305


down 4 h). The relationship between the thermal caused deviation and the thermal

load is represented by 4 different rotational speeds.

3 Modeling and simulation

There are several approaches for thermal modeling which allow the compensation of

thermo-mechanical errors like FEM models, neural networks, phenomenological

models or simplified physical models using the transient heat conduction. In this

paper, a phenomenological model and a simplified physical model are compared and

tested. Advantages and Disadvantages of these two simulation approaches compared

to FEM modeling are described in Table 1.

Table 1: Comparison between different model-approaches

Advantages Disadvantages

Simplified

physical model

(heat transfer)

- Physical model (extrapolation: unknown conditions)

- Small modeling effort

- Few measurements required

- Modeling: - Number of elements

- Geometry of elements

- Manual modeling

- Alignment of model and

measurements (Model-matching, e.g. density, heat transfer coefficient, ...)

Phenomenological

modeling

- No physical model necessary

- Only measurements necessary

- Low Uncertainties, good quality of model

- Many measurements necessary

(takes time)

- Uncertainty in unknown conditions

FEM - Physical model (extrapolation:

unknown conditions)

- If model is available from

phase of design: small modeling

effort

- Complex model (Implementation in

NC very difficult) - Alignment of model and

measurements (Model-matching, e.g.

density, heat transfer coefficient, ...)

To compare both approaches, an axis movement sequence according to Figure 4 was

measured and simulated (with the physical and the phenomenological model) for a

vertical rotary axis with a direct drive system.

Figure 4: Test sequence with varying thermal load for verification of error models

306


3.1 Phenomenological model

The structure of the phenomenological model presented in this paper can be seen in

Figure 5. It is based on three R-Test measurements at different thermal loads (at 33%,

66% and 100% of the maximum power) for parameter identification. Each

measurement returns location and positioning errors in the form of first-order lag

elements as a response to the thermal load induced into the system. As model

parameters, proportionality constants and time constants of the three measurements

are used. A simple linear interpolation between these three sampling points provides

compensation parameters for all other conditions. To consider the environmental

temperature variation, an environmental temperature variation error test (ETVE test)

was carried out over one week and implemented into the model.

Figure 5: Structure of phenomenological simulation model

Figure 6: Comparison between measurement and phenomenological simulation

Figure 6 shows the measured and computed results of two location errors for the

introduced test sequence: X0C (C-Axis movement in X-direction) and Z0T (axial

growth of machine table). For a reasonable verification, the test sequence was carried

out at different thermal loads as the measurements used for model set-up.

307

http://dict.leo.org/ende?lp=ende&p=DOKJAA&search=proportionality&trestr=0x8001

http://dict.leo.org/ende?lp=ende&p=DOKJAA&search=constant&trestr=0x8001


3.2 Physical model

The physical model presented in this paper is based on a discretization of the machine

structure in a few significant elements (Figure 7). Each of these elements represents a

part of the structure and its physical properties as mass, heat capacity, convection,

heat conduction or cooling power. The temperature in each element is assumed as

homogenous and based on the temperature distribution over all elements the

deformation can be computed. The main parameter is the current power input of the

drives of the rotary axes, which is read from the NC online via a C++ code.

Figure 7: Discretization of a tilting rotary table unit by 5 significant elements

Figure 8: Measurements and a first simulation approach with physical model

In Figure 8, first simulation results with physical model are compared to

measurements (the underlying thermal load is according to Figure 4). The figure

shows, that the basic characteristic of the location errors can be simulated very well,

but the magnitude and the transition between different proportional constant and

delay times has to be improved. A possible solution could be to use the physical

model together with a parameter identification. Using measured data, it would be

possible to identify parameters or reduce the uncertainty of estimations.

308


4 Compensation

The results of a first compensation of the location errors X0C and Z0T are shown in

Figure 9. The compensation is based on the phenomenological model presented

above. In this first approach, the compensation data was implemented via the NC

code. With the compensation, the thermo-mechanical deviation could be successfully

reduced by up to 75%.

Figure 9: Measurements with and without compensation

(based on phenomenological model)

5 Conclusion & Outlook

Two approaches for thermal modeling of location and position errors of rotary axes

of machine tools were compared. The implementation of a first phenomenological

approach showed a reduction of two different thermal location errors up to 75%. As a

next step, the physical model shall be tested and implemented into a NC. As an

extension, a parameter identification is planned. Therefore, the model shall be

adaptable easily to different machine tools or environmental conditions.

References:

[1] Mayr J. et al. (2012) Thermal issues in machine tools. Annals of the CIRP,

61/2:771-791

[2] Gebhardt M. et al. (2012) Measurement setups and -cycles for thermal

characterization, Proceedings of the 12th euspen Int. Conf., Stockholm, 1/486-489

[3] Ibaraki S., Hong C. (2012) Thermal Test for Error Maps of Rotary Axes by R-

Test, Key Engineering Materials Vols. 523-524 / pp 809-814

309


Compact Translatory Actuator with Moving Magnets and

Flexure Guide for Versatile Applications

T. Bödrich, F. Ehle, J. Lienig

Technische Universität Dresden, Institute of Electromechanical and Electronic

Design, Germany

[email protected]

Abstract

Translatory motions for small strokes up to appr. 25 mm can advantageously be real-

ised with simple single-phase linear direct drives. Compared to voice-coil actuators,

motor designs with moving permanent magnets and slotted stator winding offer high-

er forces related to winding losses and volume. Furthermore, the limited stroke allows

for utilization of stick-slip-free flexure guides. An actuator based on such a design is

presented in this paper. It is intended for position- or force-controlled operation in

small machine tools, automation and assembling.

1 Introduction

Simple single-phase electrodynamic feed units for strokes in the cm range are current-

ly being developed at Technische Universität Dresden within a German research

programme on future small machine tools [1]. Motor designs based on moving per-

manent magnets and a slotted single-phase stator winding are utilized for these actua-

tors, since the volume-based actuator constant

VP

FE

Cu

2

(1)

(F thrust force at fixed mover, PCu winding losses, V envelope volume) of moving-

magnet actuators is 2…3 times greater than that of comparable moving-coil actuators

[2]. This higher compactness is mostly due to smaller air gaps and larger cross-

sectional winding areas (and hence larger magnetomotive forces) possible with

slotted single-phase stator windings compared to moving coils [3].

A first prototype of such a feed unit of size (40 x 44 x 42) mm3 with 11 mm stroke of

a ball-guided slide, a peak force of 39 N and an embedded position control is present-

ed in [4]. Lateral magnetic attraction forces between the permanent magnets of the

310



mover and the stator as present in this cube-shaped feed unit can be avoided with an

axisymmetric actuator design, since those forces compensate along the mover cir-

cumference. This allows for a virtually stick-slip-free flexure-based guide of the

mover, making such an actuator suitable for precise positioning tasks.

2 Actuator Design

Fig. 1 and Fig. 2 show the newly developed translatory moving-magnet actuator with

flexure guide. A tubular mover with radially polarized NdFeB permanent magnets

moves translatory between two ferromagnetic stator sections in axial direction. The

outer stator contains a single-phase winding concentrically wound around the mover

magnets and the inner stator (Fig. 2). In order to effectively minimise eddy currents

during dynamic operation with a simple mechanical design, the stator components are

made of a soft-magnetic composite material rather than of radially stacked electric

sheets. Design variants and dimensioning of the magnetic circuit of those single-

phase moving-magnet actuators are outlined in [3] and [5]. The axial travel range of

the mover equals the axial width of each of the two poles of the outer stator section

(14 mm). With the chosen magnetic design the magnetic force is nearly constant

along the stroke range and proportional to the current (Fig. 3a).

Figure 1: Moving-magnet translatory

actuator with flexure guide

Figure 2: Schematic cross-sectional view

of the actuator

The net actuator force usable for actuation of loads differs from the magnetic force by

the restoring force of the flexure guide (Fig. 3b). The small total axial stiffness of the

311


flexure guide of 3.55 N/mm results in small ohmic losses of 4.4 W (at 40 °C winding

temperature) in both mover end positions due to deflection of the guide. With careful,

mostly FEA-based design of the flexures it was possible to realise a large stroke of

±7 mm = 14 mm with an outer diameter of each flexure of only 59 mm, resulting in a

compact overall actuator design.

Figure 3: Force-position-current characteristic of the actuator a) without and

b) with the restoring force of the flexure guide

3 Preliminary Technical Data

Features of the developed moving-magnet actuator with flexure guide are:

travel range 14 mm,

continuous magnetic force 44 N (without restoring force of the flexure guide;

higher continuous force possible with higher wire insulation class),

peak force 112 N (without restoring force of the flexure guide, see Fig. 3),

compact magnetic circuit 67 mm, axial length 32 mm (total axial length

72 mm due to space for deflection of the flexure guide),

volume-related actuator constant incl. space for flexure guide 0.62 N2/(W cm3),

of magnetic subsystem only excl. space for flexure guide 1.32 N2/(W cm3) [2],

mover mass 0.085 kg,

inductance 179 mH, electrical time constant of the slotted winding 16 ms,

with 48 VDC supply voltage during position-controlled stroke maximum veloci-

ty of 1.7 m/s and peak acceleration of 34 g to be expected.

312


The control electronics already developed for state space position control of the

above-mentioned cube-shaped feed unit [4] is currently redesigned and integrated

into the newly developed cylindrical actuator, together with a low-cost optical incre-

mental position sensor. Flatness-based control of the mover position or force resp. is

currently implemented in this electronics. A positioning accuracy of 1 µm or better is

aimed for.

4 Conclusions and Outlook

The developed moving-magnet actuator features a linear force characteristic, high

force density and good dynamic behaviour. The integrated flexure guide makes it

suitable for precise positioning tasks, especially with the position sensor and embed-

ded control electronics currently being integrated. With latter components the actua-

tor can become a compact and cost-efficient drive unit for versatile applications in

small machine tools, automation and assembling.

5 Acknowledgements

The authors would like to thank the German Research Foundation (Deutsche For-

schungsgemeinschaft - DFG) for funding of the presented work within the Priority

Programme SPP 1476 "Small machine tools for small work pieces".

References:

[1] Wulfsberg, J. P.; Grimske, S.; Kong, N.: Kleine Werkzeugmaschinen für kleine

Werkstücke. wt Werkstattstechnik online 100 (2010) 11/12, pp. 886-891

[2] Bödrich, T.; Süßenbecker, M.; Ehle, F.; Lienig, J.: Kompakte einphasige

Lineardirektantriebsmodule für kleine Verfahrwege. ant Journal 1/2013, pp.

16-21

[3] Bödrich, T.: Modellbasierter Vergleich einphasiger permanentmagneterregter

translatorischer Wandler (Model-Based Comparison of Single-Phase, Perma-

nent-Magnetically Excited Translatory Converters). ETG-Fachbericht 118+119,

Berlin, Offenbach: VDE-Verlag 2009, pp. 85-90

[4] Bödrich, T.; Süßenbecker, M.; Lienig, J.: Electrodynamic Feed Units for Small

Machine Tools. Proc. of 12th euspen Int. Conf., Stockholm, June 4-8, 2012,

Vol. 1, pp. 519-522

[5] Jack.; A. G.; Al-Otaibi, Z. S.; Persson, M.: Alternative Designs for Oscillating

Linear Single Phase Permanent Magnet Motors Using Soft Magnetic Compo-

sites. Proc. of ICEMS 2006 Int. Conf. on Electrical Machines and Systems,

Nov. 20-23, 2006, Nagasaki, Japan, Paper ID DS4F2-07

313


Displacement of a 6-DOF inchworm-based parallel

kinematic stage

A. Torii, R. Kamiya, K. Doki, A. Ueda

Dept. of Electrical and Electronics Eng., Aichi Institute of Technology, Japan

[email protected]

Abstract

A novel inchworm-based positioning stage with large motion ranges for microscopes

and machine tools is presented. Performance of the stage related to control signal is

described. We discuss parasitic motions, which are the displacement orthogonally to

the motion plane. The control signal, which reduces the parasitic motions, is

determined. The result obtained in this paper is useful for continuous path control and

the position and orientation control of the stage.

1 Introduction

Recent machine tools require high capability for the ultrafine machining machines

and for micro/nano positioning. In order to achieve high resolution, piezoelectric

actuators (piezos) are used. Multi-axis machining of surfaces requires not only

precision position control but also precision orientation control. We described the

concept and design of a six-degree-of-freedom (6-DOF) micro parallel kinematic

stage for multi-axis positioning[1]. The stage realized about 10 nm linear

displacement. Since the stage was driven by the principle of an inchworm, the stage

showed the parasitic motion, which is the displacement orthogonally to the motion

plane. In this paper, we discuss the control signal, which reduces the parasitic

motions of the stage.

2 Structure of Six-DOF Stage

Figure 1 shows the 6-DOF parallel kinematic stage. Six stacked-type piezos and six

electromagnets are used. The proposed 6-DOF stage is based on the hexapod

structure. Six metal parts bond the piezos in 109.5 degrees. The stage does not have a

fixed base, therefore it has a large motion range on a surface. Three electromagnets A,

C, E touching on the base can connect/disconnect the stage and the base. The other

314


three electromagnets B, D, F supporting the platform can connect/disconnect the

stage and the platform. The non-excited electromagnets move sequentially by the

deformation of the piezos. The stage moves in six directions (x, y, z, x, y and z),

realizes sub-micron preciseness, and has an unlimited working area.

The piezo, which is 10 mm in length, deforms 5 m when 100 VDC is applied. The

electromagnetic force is about 5 N when 10 V is applied. The electromagnets and

piezos are controlled synchronously, and they rotate and tilt a hemispherical platform.

The size of the stage is about 50 mm by 50 mm and 50 mm in height, which depends

on the dimensions of the piezos and electromagnets. A platform used in the

experiment is a 100 mm square iron flat plate with 1 mm thick.

a. b.

Figure 1: (a) Photo of the 6-DOF stage and (b) schematic diagram of the stage.

3 Control of Six-DOF Stage

3.1 Inchworm motion

In our previous work, the control signal for the stage is based on the principle of an

inchworm. Although the principle of an inchworm helps overcome the problem of

poor working range, it causes the parasitic displacement. While five out of six

electromagnets are excited, the other electromagnet that is not excited moves by the

deformation of the piezos. The piezos deform in the longitudinal direction, and thrust

the electromagnet. The electromagnet A, C, E move on the base, and the

electromagnet B, D, F move on the platform.

In Figure 2(a), voltages applied to the electromagnets and piezos are illustrated. One

control cycle is 1 s. The stage connected with a platform by the electromagnets B, D,

F moves in horizontal y-direction. Non-excited electromagnets A, C, E sequentially

315


move by the deformation of two piezos. The voltage applied to the piezo is ramp

input and the maximum voltage is 100 V. The constant value of the maximum

voltage applied to the piezo causes the parasitic motion of the stage. Figure 2(b)

shows the horizontal y-displacement and parasitic vertical z-displacement of the

platform. Although the horizontal displacement for 10 cycles is about 50 m, the

parasitic displacement in z-direction is 7.5 m.

a. b.

Figure 2: (a) control signals applied to electromagnets and piezos for horizontal

displacement and (b) horizontal y-displacement and vertical z-displacement.

3.2 Control Signal Using Inverse Kinematics

The control signal proposed in this paper is determined by the inverse kinematics.

The deformations of the piezos a, b, ..., f are expressed by dLi (i=1, 2, ... 6),

respectively. The deformation of the piezo dL={dL1, dL2, dL3, dL4, dL5, dL6}T is

expressed by

dL=J dq, (1)

where dq={dx, dy, dz, d x, d y, d z}T denotes the minute displacement of the

platform, and J denotes a 6x6 Jacobi matrix which is obtained from the geometrical

consideration. The voltage applied to the piezo, which is ramp input, is determined by

equation (1). In our experimental setup, a personal computer generates a control

signal, which is applied, to the piezo through a voltage amplifier.

Figure 3(a) shows the modified control signals which are used for horizontal y-

displacement of the stage. The deformation of the piezo is calculated by the inverse

kinematics. The voltage applied to the piezo is determined under the assumption that

the deformation of the piezo is proportional to the applied voltage, although the

maximum voltage applied to the piezo is 100 V. The speed of the piezo deformation

316


varies according to the voltage applied. Figure 3(b) shows the horizontal y-

displacement and vertical z-displacement. By changing the control sequence and the

amplitude of the voltage applied to the piezo, the parasitic z-displacement is reduced

to 2.7 m, although the horizontal y-displacement for 10 cycles is also reduced to 32

m. The fine motion is obtained by the inverse kinematics. The desired position and

orientation of the stage determines control signals of the piezos by equation (1).

a. b.

Figure 3: (a) modified control signals applied to electromagnets and piezos, and (b)

horizontal y-displacement and vertical z-displacement.

4 Summary

Control signals, which are applied to the positioning stage, are discussed. We change

the sequence and voltage of the control signals. Desired position and orientation of

the stage determines the control signal of the piezo. In the experiment, the stage

motion in horizontal y-direction and parasitic vertical z-direction is measured. The

parasitic displacement is reduced by considering the inverse kinematics of the stage.

Acknowledgements:

This project was financially supported by the Japan Society for the Promotion of

Science (JSPS), “Grant-in Aid for Scientific Research (C), No. 23560302”.

References:

[1] A. Torii et al., A six-degree-of-freedom micro parallel kinematic stage for multi-

axis positioning, 12th International conference of the euspen, Stockholm, Sweden,

vol. 1, pp. 543-546, 2012

317


Increased Productivity due to Jerk-decoupled Feed Axes of

the 5-Axes Milling Machine “Neximo”

B. Denkena, K. Litwinski, O. Gümmer

Institute of Production Engineering and Machine Tools (IFW), Leibniz Universitaet

Hannover, Germany

[email protected]

Abstract

With the development of highly dynamic machine tools, flexibilities of the machine

structure represent a major challenge in machine tool design. High machining forces

and accelerations induce static and dynamic deflections of the machine frame which

affect the accuracy of the machine negatively. To master this conflictive aim of high

dynamic feed drives and simultaneously increased machining accuracy, the IFW

developed the 5-axis machine tool prototype "Neximo" with jerk-decoupled x- and y-

axes. This paper presents the results of the performed vibration measurements and the

achievable increased productivity in machining processes by higher jerk limits due to

the applied jerk decoupling technology.

1 Introduction and Motivation

Dynamic feed drives provide high acceleration gradients which also lead to an intense

vibration excitation of the machine frame. This conflict is traditionally met by a

severe limitation of the reference value regarding the jerk and therefore of the

effective dynamics of the drive [1]. Thus, the reduction or avoidance of the machine

frame excitation without jerk limitation is a relevant research topic. For this purpose

different approaches have been developed in terms of the impulse- or jerk-decoupling

technology, the impulse compensation and the trajectory shaping [1]. Cross

references to older developed systems are given in the presented recent references.

Table 1: Resulting time saving at different jerk limits for a position step of 200 mm

Jerk limit positioning time time saving

50 0.504 s -26.0 %

100 0.4 s 0 %

250 0.2947 s 26.32 %

500 0.244 s 39.03 %

1,000 0.221 s 44.78 %

2,000 0.2102 s 47.47 %

318



Figure 1: Position trajectory at different jerk limits

Exemplarily considering the resulting positioning time for a position step of 200 mm

of a machine axis with a maximum speed of 120 m/min and a maximum acceleration

of 20 m/s² (see Figure 1 and Table 1) clearly shows the potential of increasing the

jerk limitation. Compared with a conventional standard maximum jerk of 100 m/s³, a

tenfold to 1,000 m/s³ leads to reduction of about 45% in positioning time.

2 Machine tool “Neximo”

For increasing the jerk limit without increasing the vibration excitation of the

machine frame, a new 5-axis machine tool prototype with integrated innovative jerk-

decoupling technology in the x- and y-axis (see Figure 2) was developed at the IFW

[2]. This prototype enables the analysis of this technology in 5-axes milling

processes. Additionally, the z-axis is equipped with an active magnetic guidance for

the compensation for static and dynamic errors and enhancement of the machining

accuracy. The jerk-decoupling technology is based on a movable secondary part of

the linear direct drive, which is connected to the machine frame by independently

adjustable spring-damper-elements [3]. This arrangement enables a mechanical low-

pass filtering of the dynamic drive forces, so that the vibration excitation of the

structural machine frame modes can be significantly reduced compared to non-

decoupled feed axes. As spring-elements pneumatic muscles are used to vary the

stiffness by the pneumatic pressure and the damping can be adjusted by using a

magnetic damping unit. The movable secondary part is guided through the new and

Güv/70973 ©IFW

319


patented relative guidance [3]. Hereby, the investment costs and the friction losses of

the needed additional guidance for the secondary part are clearly reduced.

Figure 2: Jerk-decoupled machine tool “Neximo”

3 Increase in Productivity

For the analysis of the effectiveness of this technology vibration measurements of the

x- and y-axis of the “Neximo” machine tool are carried out. The vibration is

measured on the basis of the position signal of the linear scale of the feed axes, which

is comparable to the relative vibration between tooltip and workpiece, as comparative

measurements have shown. Additionally, the machine frame vibration is externally

measured using a laser vibrometer. The experiments show a reduced vibration

response with active jerk-decoupling (with JDC) in comparison to the measurements

with clamped jerk-decoupling slide (without JDC).

Figure 3: Vibration measurement without and with jerk-decoupling (jerk: 250 m/s³)

Güv/70974 ©IFW

Güv/70975 ©IFW

320


Exemplarily, in Figure 3 the comparison of the vibration measurements of the x-axis

for a positioning step of 10 mm with rapid traverse and a jerk limit of 250 m/s³ are

presented. Hence, the effectivity of the jerk-decoupling technology in a milling

machine was verified and the jerk limitation could thereby be increased in the x-axis

from 50 to 200 m/s³ and in the y-axis from 100 to 500 m/s³.

In addition, two reference workpieces of the German NC-society were manufactured

with and without the jerk-decoupling technology. The results show a reduced

processing time for the “Test Workpiece for the 5-Axis Simultaneous Milling

Machining” of 3.2 %. The comparison of the machining time for the 3-axes “Test

Workpiece for High Speed Cutting (HSC)” is given in Table 2. Due to higher jerk

limits an increase in productivity of 1.8 % for the roughing process and 4.3 % for the

finishing process are achieved without negatively influencing the machining

accuracy, which has been confirmed by measurements on a CMM.

Table 2: Machining time of the 3-axes workpiece without and with jerk-decoupling

jerk limit

x-axis y-axis z-axis

machining time

roughing finishing

without JDC 50 m/s³ 100 m/s³ 100 m/s³ 2:43 min 1:34 min

with JDC 200 m/s³ 500 m/s³ 100 m/s³ 2:40 min 1:30 min

4 Conclusion

The carried out oscillation measurements of the new machine tool prototype

“Neximo” confirmed the reduction of the machine frame vibrations by using the jerk-

decoupling technology. Thus, the jerk limit and productivity of machine tools can be

increased without negatively influencing the machining accuracy, whereupon the

achievable percentage enhancement severely depends on the workpiece. In addition,

the tool changing time is clearly reduced, which is currently not taken into account.

References:

[1] Altintas, Y.; Verl, A.; Brecher, C.; Uriarte, L.; Pritschow, G.: “Machine tool

feed drives”, CIRP Annals - Manufacturing Technology, Vol. 60, Issue 2, pp.

779–796, 2011.

[2] Denkena, B.; Möhring, H.-C.; Gümmer, O.: “Hochdynamische ruckentkoppelte

Werkzeugmaschine, wt Werkstattstechnik online, Vol. 100 Heft 1/2, pp. 99–104,

2010.

[3] Denkena, B.; Hesse, P.; Gümmer, O.: “Energy optimized jerk-decoupling

technology for translatory feed axes”, CIRP Annals - Manufacturing Tech-

nology, Vol. 58, Issue 1, pp. 339–342, 2009.

321


Design and optimization of a 3-DOF planar MEMS Stage

with integrated thermal position sensors

B. Krijnen1,2, K. R. Swinkels1,2, D. M. Brouwer1,2, J. L. Herder2 1DEMCON Advanced Mechatronics, The Netherlands 2Mechanical Automation & Mechatronics, University of Twente, The Netherlands

[email protected]

Abstract

This work presents the design and optimization of a large stroke planar positioning

stage in a single-mask MEMS fabrication process. Electrostatic comb-drive actuators

were used to control the position and rotation of the 3-DOF stage. Thermal

displacement sensors are integrated to provide feedback. Simulations show that we

are able to reach a +/-120 m range of motion and +/-30 degrees of rotation.

Preliminary measurements were performed which validated our models.

1 Introduction

MEMS positioning stages can be used in a variety of applications, such as micro-

mirror manipulation, scanning probe microscopy and probe based data storage.

Existing multi-DOF (Degree-Of-Freedom) stages often lack integrated position

sensing [1], use complicated fabrication schemes and assembly [2, 3], or offer

relatively small stroke [4]. The latter uses thermal actuators which is unfavourable for

fast and accurate positioning. In this work we describe the design, optimization, and

fabrication of a large stroke planar positioning stage with integrated displacement

sensors for feedback control in a simple, single-mask fabrication process.

2 System design

The system consists of a 3-DOF stage that is connected with three single-DOF

shuttles using leaf springs. An overview of one of the shuttles and the eccentric

connection to the stage is given in Figure 1 and Figure 3. The single-DOF shuttles are

actuated by electrostatic comb-drive actuators [5]. The electrostatic field of the comb-

drive actuators give rise to a negative lateral stiffness, which can lead to instability

(pull-in) if the lateral mechanical stiffness of the flexure mechanism is not sufficient.

322



The flexible multibody analysis software SPACAR [6] was used to numerically

analyse the mechanical stiffness of the shuttles in actuation and lateral direction as a

function of the displacement. The data from SPACAR is used to calculate if pull-in

occurs for any stage position and rotation. In this way the work space of the 3-DOF

stage is determined.

Figure 1: The figure gives an overview of one of the single-DOF shuttles together

with the connection to the stage. The arrow indicates the DOF of the shuttle.

For feedback control of the stage, thermal displacement sensors are integrated in the

design [7]. The temperature of the heaters changes due to a varying overlap with the

'cold' shuttle. The resulting change in electrical resistance was measured and results

in a position resolution of 4nm at a bandwidth of 30Hz. To control the stage position

using the displacement sensors on the shuttles, a geometric transfer function

(mapping) between the shuttle positions and the stage position and orientation was

developed. Simulations showed that with a double integrator control scheme (PII) we

are able to control the position of the stage with an accuracy of 22nm and 0.17mrad.

3 Optimization

The work space of the 3-DOF stage is restricted by pull-in of the single-DOF shuttles

in push as well as pull direction. This results in a hexagonal work space, which is

given for several rotations in Figure 2 (left). The size of the stage (eccentricity), the

length of the leaf springs towards the stage, and the point of attachment of the leaf

springs to the stage and shuttles are varied to optimize the range of motion of the

323


stage (the largest circle in the hexagonal work space). A stage eccentricity of 110 m

leads to a maximum range of motion at zero rotation of +/-120 m, as can be seen in

Figure 2 (right). For a rotation of 10° the range of motion decreases to +/-75 m.

Figure 2: The work space of the 3-DOF stage is restricted by pull-in of the single-

DOF shuttles in push as well as pull direction. This results in a hexagonal work

space (left). The range of motion in micrometers is a function of the stage

eccentricity and has an optimum of +/-120 m at zero rotation (right).

3 Fabrication and results

The complete system was designed to be integrated in a standard fabrication process

based on a silicon-on-insulator wafer. Deep Reactive-Ion Etching (DRIE) is used to

anisotropically etch high aspect ratio trenches. Thin structures are released from the

handle wafer by isotropic VHF etching of the buried oxide layer, while wide

structures stay anchored.

In spite of stiction in most of the fabricated devices, a number of measurements could

be performed to validate our models. The voltage that leads to pull-in was measured

for several stage designs and indicated that our model is correct within 10% with

respect to the simulated pull-in voltages.

4 Conclusion

A 3-DOF planar MEMS positioning stage with integrated position sensors was

designed and fabricated in a single-mask fabrication process (Figure 3). The stage is

able to reach a +/-120 m range of motion and +/-30 degrees of rotation.

324


Figure 3: The figure shows a scanning electron microscope image of a fabricated 3-

DOF stage together with the eccentric connection towards the three shuttles. Parts of

the flexure mechanisms of the single-DOF shuttles are also visible.

References

[1] D. Mukhopadhyay, et al. A SOI-MEMS-based 3-DOF planar parallel-kinematics

nanopositioning stage. Sensors and Actuators A, 174(1):340–351, 2008.

[2] L. Sun, et al. A silicon integrated micro nano-positioning XY-stage for nano-

manipulation. J. of Micromechanics and Microengineering, 18(12):125004, 2008.

[3] E. Eleftheriou, et al. Millipede—A MEMS-Based Scanning-Probe Data-Storage

System. IEEE Transactions on Magnetics, 39(2):938-945, 2003.

[4] L.L. Chu and Y.B. Gianchandani. A micromachined 2D positioner with

electrothermal actuation and sub-nanometer capacitive sensing. J. of

Micromechanics and Microengineering, 13(2):279-285, 2003

[5] R. Legtenberg, et al. Comb-drive actuators for large displacements. J. of

Micromechanics and Microengineering, 6(3):320-329, 1996.

[6] J.B. Jonker and J.P. Meijaard. SPACAR - Computer Program for Dynamic

Analysis of Flexible Spatial Mechanisms and Manipulators. In W. Schiehlen, editor,

Multibody Systems Handbook, pages 123–143. Springer, 1990.

[7] B. Krijnen, et al. A single-mask thermal displacement sensor in mems. J. of

Micromechanics and Microengineering, 21(7):074007, 2011.

325


Sensorless monitoring of machining torque on tilting

platform driven by hybrid actuator

H. Yoshioka1, M. Hayashi2, H. Sawano1, H. Shinno1 1Tokyo Institute of Technology, Japan 2The University of Tokyo, Japan

[email protected]

Abstract

This paper presents a sensorless monitoring method of machining torque on a tilting

platform driven by a hybrid actuator. Sensorless monitoring function can estimate

machining torque without additional sensor devices. Performance evaluation results

of basic machining tests confirm that the developed system has a capability to

monitor machining torque on tilting platform driven by the hybrid actuator.

1 Introduction

Multi-axis controlled precision machine tools have been developed to produce precise

parts with complex and freeform geometries. In order to realize freeform nano-

machining, it is necessary to develop high performance rotary and tilting platforms.

The tilting platform driven by a hybrid actuator that was successfully integrated a

pneumatic actuator with an electric actuator has been developed for nano-machining

[1]. To realize higher performance machining, the machining force during process

should be monitored to control suitable machining condition. However, it is difficult

to install monitoring sensors into a rotary axis because it needs hardwiring.

This paper describes a sensorless monitoring method of machining torque on a tilting

platform driven by a hybrid actuator. After installing a sensorless monitoring function

into a controller, performance evaluation was carried out through actual machining

experiments.

2 Sensorless monitoring function for tilting platform

2.1 Tilting platform driven by hybrid actuator

Figure 1 shows a developed tilting platform for precision machining driven by a

hybrid rotary actuator [1]. The hybrid actuator consists of a pneumatic actuator at the

326


center of driving shaft and voice coil motors at both ends of the shaft. They are

controlled to generate quick and accurate torque by complementing their

characteristics each other. Because the platform is supported by aerostatic bearing it

can move without friction. The table is installed in parallel with the driving shaft.

This structure provides wide working area that is almost equal to the footprint.

Figure 1: Structural configuration of the developed tilting platform

2.2 Sensorless monitoring function for the platform

Sensorless monitoring function can estimate machining torque using both the output

of controller and the measured position based on a disturbance observer [2-3].

Because there is no friction in the developed platform, rotary motion can be

expressed by the following equation:

dKTJ ref (1)

where is angle, J is inertia of the platform, K is torque constant, d is disturbance,

and Tref is output of controller, respectively. Therefore, a discretized equation of the

system is expressed as the following equation using state variable vector x defined:

Tdx (2)

][][

][][]1[

iiy

iTii ref

xC

HxGx (3)

001,

0

2

,

100

10

2122

CHG JKT

JKT

JT

JTTs

s

s

s

s

(4)

327


where Ts is the sampling period of the controller. Thus, an observer for the system is

represented by the following equation:

][][][ˆ

][ˆ][][][ˆ]1[ˆ

iyiTi

iiyiTii

ref

ref

ee

e

KHxCKG

xCKHxGx (5)

where x is the estimated state variable and Ke is the feedback gain of observer which

is determined to estimate x stably. In this study, roots of CKG e were set at 100Hz.

The designed sensorless monitoring function was installed into the controller.

3 Evaluation of the monitoring performance

3.1 Experimental setup for evaluation

Figure 2 shows the experimental setup for performance evaluation. The tilting

platform was controlled at 0 degree by the positioning controller and a workpiece was

fixed on the platform. Machining motion was provided by a feed axis with a diamond

tool fixed on a dynamometer. Therefore, the estimated machining torque by a

sensorless monitoring function in the controller was compared with the output of

dynamometer. Machining conditions are shown in Table 1.

Table 1: Machining conditions

Tool Single crystal

diamond (R=0.2)

Depth of cut 5 m, 10 m

Cutting speed 20mm/s

Cutting fluid Dry

Workpiece Brass

Figure 2: Experimental setup

3.2 Monitoring results

Figure 3 shows the measured torque by the dynamometer and the estimated

machining torque by the sensorless monitoring function. Because there is no friction

in the developed tilting platform, it is easy to make a system model and to identify

system parameters accurately. Therefore, the estimated machining torque shows good

agreement with the measured torque in both cases.

328


(a) Depth of cut 5 m (b) Depth of cut 10 m

Figure 3: Monitored machining torque during process

4 Conclusions

This paper presents the sensorless monitoring method of machining torque for tilting

platform driven by the hybrid actuator. Evaluation results of actual machining

experiments confirmed that the designed monitoring function can monitor the torque

during machining without any additional sensor devices.

Acknowledgements

This research project was financially supported by the Japan Society for the

Promotion of Science (JSPS) Grant-in-Aid for Scientific Research (S) No.24226004.

References:

[1] Hayashi,M., et al., A hybrid actuator-driven compact tilting motion table system

for multi-axis ultraprecision machine tool, Proc. 11th euspen Int. Conf., Vol.2,

2011, pp.19-22.

[2] Shinno, H., et al., Sensor-less monitoring of cutting force during ultraprecision

machining, CIRP Annals, v 52, n 1, 2003, pp. 303-306.

[3] Tanaka, H., et al., Torque sensor-less tactile control of electrorheological passive

actuators, Int. Workshop Adv. Motion Cont. AMC, 2010, pp. 331-336.

329


Self-tuning dynamic vibration absorber for machine tool

chatter suppression

G. Aguirre1, M. Gorostiaga1, T. Porchez2, J. Muñoa1 1IK4-IDEKO, Spain 2CEDRAT TECHNOLOGIES, France

[email protected]

Abstract

The current trend in machine tool design aims at stiffer machines with lower

influence of friction, leading to faster and more precise machines. However, this is at

the expense of reducing the machine damping, which is mainly produced by friction,

and thus increasing the risk of suffering from a self-excited vibration named chatter,

which limits the productivity of the process. Dynamic vibration absorbers (DVAs)

offer a relatively simple and low cost solution to reduce chatter appearance risk by

adding damping to machine tools. The proper tuning of the dynamic characteristics of

the damper to the machine/process dynamics is the key for productivity improvement.

A DVA which can detect its optimal tuning frequency and adapt its dynamics

accordingly is proposed here. The main design and working principles of this damper,

and the improvement of the machining conditions allowed by the damper will be

demonstrated by real milling experiments.

1 Design of Dynamic Vibration Absorber

Dynamic vibration absorbers consist of a mass connected to the machine with a

certain stiffness and damping, so that its resonance frequency is tuned to the

frequency of the machine mode leading to chatter, by adding damping to it and

allowing higher cutting depths [1]. Tunable DVAs are needed so that their dynamic

characteristics can be adapted to a range of machine and process conditions.

The main challenges that need to be addressed are how to obtain the desired stiffness

and damping values, and how to detect the frequency to which it needs to be tuned.

The solution proposed here allows independent stiffness and damping tuning, with

online stiffness tuning capability, with a highly repetitive and linear behaviour,

compared to typical solutions based on elastomers.

330



In-process optimal tuning frequency detection is proposed too, based on processing

an accelerometer signal, thus without need of an experimental modal analysis, which

requires specialized equipment and personnel, and does not consider changes during

machine operation (e.g. different workpiece mass, position dependent stiffness, etc.).

1.1 Variable stiffness spring

The stiffness of the DVA proposed here can be varied

thanks to a variable stiffness spring controlled by a

rotary stepper motor (see Figure 4a). Within 90º

rotation, the stiffness of the spring changes between

two values, easily defined at design stage through a

and b parameters (see Fig. 1), providing a repetitive

and linear stiffness tuning.

1.2 Eddy current damping tuning

Eddy current damping is generated by the relative motion of copper plates within a

magnetic field generated by magnets, and it provides a close-to-viscous damping

effect. Since it generates no stiffness, and the spring presented above provides little

damping to the system, the stiffness and damping of the DVA can be tuned

independently, which is a great advantage over other typical solutions such as

elastomers.

Figure 2: Magnet configuration and orientation, design and prototype

1.3 Self-tuning strategy

Optimal performance can be achieved if the tuning frequency can be calculated

online during the machining process. A chatter detection algorithm is proposed here,

which based on the information provided by an accelerometer placed on the DVA,

detects whether chatter is being generated and at which frequency, and tunes the

Figure 1: Rotary spring

331


damper accordingly by commanding the stepper motor to move the rotary spring

according to a calibrated angular position – resonance frequency relationship.


Machining tests have been carried out on a SORALUCE

milling machine, mounting the workpiece on a flexible

fixture for development and evaluation purposes. This

fixture provides a dynamic response of the machine with

a clear and isolated resonance mode, prone to suffer from

chatter, and thus of help to avoid other disturbing effects,

such us modes at similar frequencies, which would

difficult evaluation of the performance of the semi-active

damper presented here. Anyway, this is still a realistic

test case comparable to many industrial cases.

A DVA prototype has been built to meet the requirements

of this test bench, which shows a critical mode at 94 Hz and 150 kg modal mass.

With a moving mass of 7 kg, the DVA can change its main resonance frequency

between 65 Hz and 105 Hz, providing an estimated 800 Ns/m damping, values which

are in range with the optimal [1]. As it can be seen on Figure 4, the moving mass is

formed by the four magnet racks, with the copper plates fixed to the frame, providing

thus a very compact system. An accelerometer is placed on the frame to measure

machine vibration.

Figure 4: DVA component a) rotary spring and motor b) magnets c) assembly

3 Machining results

By performing an experimental modal analysis of the machine, the stability lobes of

the cutting process have been calculated, in order to show the maximum cutting depth

that can be achieved without the damper. A number of machining tests show the

Figure 3: Machine tool

and DVA setup

332


validity of this prediction. The DVA was placed on the machine next, and it was

tuned automatically during the cutting process, without using the information from

the modal analysis. The process was found to be stable up to the maximum cutting

depth defined by the tool (5 mm), compared to the unstable conditions with 1 mm

depth without damper. A time simulation of the cutting and tuning process predicts

much higher stable cutting depths, but they cannot be reached due to tool limitations.

Figure 5: Machining test and simulation results, with and without DVA.

In Figure 6, the part that was used for the machining tests is shown. Clear chatter

vibration marks can be seen without damper and 3mm depth, while a smooth surface

is produced using the DVA with 5 mm depth, for several spindle speeds.

Figure 6: Workpiece surface: a) no DVA, 3 mm depth b) with DVA, 5 mm depth

4 Conclusions

These results demonstrate the effectiveness of the self-tuning DVA principle

presented here. In real applications, productivity improvement will not be so high, but

it will outperform existing DVAs by providing a low cost solution that does not

require a previous experimental modal analysis and that works in close-to-optimal

conditions even when process dynamics change during operation.

References:

[1] N. D. Sims, “Vibration absorbers for chatter suppression: A new analytical

tuning methodology,” Journal of Sound and Vibration, vol. 301, no. 3–5, pp. 592–

607, April 2007.

333


Design and control of a through wall 450 mm vacuum

compatible wafer stage

D. Laro1, E. Boots2, J. van Eijk2,3, L. Sanders1

MI-Partners, The Netherlands1

TU Delft, The Netherlands2

MICE BV, The Netherlands3

[email protected]

Abstract

High precision machines such as EUV wafer scanners and E-beam measurement

systems require a high vacuum level. Contamination of this vacuum due to moving

cables and bearings of the positioning stages within are an issue. An inverted planar

motor solves this contamination issue but leads to a complex system due to position

dependant commutation and a large number of coils [1]. Therefore an alternative

stage design is made at MI-Partners (see Figure 1) which has a low degree of

complexity and does not cause contamination of the vacuum. In this concept a

separation has been made between two vacuum levels: a clean/precision vacuum and

a non-precision/dirty vacuum. The separation between the two is realized by a wall.

The design uses a Short Stroke-Long Stroke (SS-LS) stage configuration where the

SS stage makes its actuation forces through the wall. The precision vacuum contains

the SS chuck carrying a wafer for manufacturing or inspection. In the non-precision

vacuum a conventional stacked LS x-y stage is placed. The function of this XY stage

is to enable a larger stroke for the short stroke system. The vacuum underneath the

separator plate is only required to minimize loads on the wall due to the pressure

difference over a large area. In this paper the design of a demonstrator and its control

architecture is described.

1 Through Wall Actuators

The through wall stage concept requires the development of new actuators due to the

large air gap introduced by the wall. To minimize the complexity of the actuators, the

functionality of suspending and propelling the SS stage is split into two separate

334



actuators units: an in-plane actuator and an active magnetic gravity compensator

(Figure 2).

1.1 Active Magnetic Gravity Compensator

The active magnetic gravity compensator holds the weight of the stage using passive

magnets and enables actuation using a coil on the LS side. The magnets minimize

power consumption for carrying the weight, but potentially adds actuator stiffness. As

the long stroke stage is envisioned to be a conventional stage without high

requirements, the transfer of disturbances and therefore actuator stiffness should be

minimized. To realize this, a low stiffness gravity compensator was developed using

a circular magnet with a hole in the center. The design principle of the zero stiffness

effect is explained in Figure 2.

1.2 In-plane Actuator

The in-plane actuator is based on a Halbach magnet arrangement. The Halbach array

ensures the magnetic fields is “pushed” towards the LS stage (Figure 2). The design

was optimized to minimize position dependency; it can be operated without the need

of commutation.

Figure 1 Through wall vacuum wafer stage design, showing the separation between

a precision/clean and a non-precision/dirty vacuum compartment, as well as the

Long Stroke Short Stroke stage configuration.

335


In-plane motor

Active Gravity Compensator

Figure 2 Custom motor designs for In-plane and out of plane actuation.

Figure 3 Working Principle of Active Magnetic Gravity Compensator. The ring

shaped magnet with force F3 can be seen as the superposition of a larger magnet

(force F1) with opposite polarisation and a smaller magnet with attractive

polarization (force F2), resulting in zero stiffness behaviour.

2 Design and Control Integration

To demonstrate the potential of the through wall concept a demonstrator has been

realized. The goal of the demonstrator is not to reach nm-precision but to show the

concept. The design consists of a 450 mm Short Stroke wafer chuck. This chuck is

actuated using four in plane actuators and four active gravity compensators. This

leads to an over-actuated system. The advantage of over-actuation is reduced

excitation of the torsional mode, which typically limits performance of large chucks

[2], see Figure 8. The stage is measured in 6 DoF with respect to a metrology frame

and is controlled using a decoupled controller around the SS’s center of gravity. A

bandwidth of 100 Hz has been achieved on all directions. The long stroke stage tracks

the SS stroke stage. The long stroke is actuated by a spindle drive. To minimize

visibility of spindle stage dynamics through the reaction path of the short stroke

336


stage, the metrology frame is isolated at 10 Hz using passive rubbers. Currently the

design has only one long stroke direction, but a stacked XY stage can easily be added

to the system. A recent development has been a wireless energy system towards the

SS stage.

Metrology Frame

Short Stroke: 450 mm chuck

Long Stroke

Wall

Force Frame

Figure 4 Mechanical layout of the system (left), image of the demonstrator (right).

Figure 5 Actuator layout of SS chuck (left). FRF of Short stroke stage in Z direction

(right). Torsional mode at 300 Hz is only mildly excited because of over-actuation.

References:

[1] J.W. Jansen. “Magnetically levitated planar actuator with moving magnets:

Electromechanical analysis and design”, PhD thesis, TU Eindhoven , pp. 5

[2] D. Laro, R. Boshuizen, J. van Eijk. “Design and control of over-actuated

lightweight 450 mm wafer chuck”, ASPE control topical meeting 2010, pp. 141-144.

* This research is conducted within the XTreme Motion project which is part of the

Pieken in de Delta program of the Dutch Ministry of Economic Affairs

337


Driving a Femtosecond Machined Tactile Scanning Probe

Stage in the 100 µm Range

D. F. Vles, F. G. A. Homburg

Eindhoven University of Technology, Netherlands

[email protected]

Abstract

We propose a novel meso-sized electromagnetic actuator design for driving a

femtosecond machined tactile scanning probe stage in the 100 m range. The final

design uses a static flat single-turn coil with heat fins, with which thermal limitations

usually faced in high precision actuation are overcome. High current densities up to

108 A/m2 can be reached. On a PCB, we easily manufacture a first prototype, with

which we predict to achieve a (constant) 5 mN force over a 200 m range.

1 Introduction

Systems built for actuation on the micro-scale are of high interest in the field of

optics and high precision 3D metrology systems. Due to complex interplay of the

different scaling laws at the meso-scale, such actuators are usually micro- or macro-

sized [1]. Continuing on the research of Boustheen et al [2], we investigate the effects

of scaling laws on the performance limits of high force (1 mN), high stroke (100 μm)

meso-sized (0.5 – 2 mm3) actuators, i.e. electrostatic, electromagnetic and

piezoelectric actuators. The best performing actuation method is selected, optimized

and fabricated for the actuation of a 3 degree-of-freedom (3DOF) stage, suitable for

ultra precision 3D tactile scanning probes.

2 Selection of Actuation Principle

The choice of actuation principle for a certain application is usually based on

databases and determined by interpolating (using scaling laws) between performance

characteristics of documented (commercial) actuators. Such interpolation procedures

and the few available meso-sized actuators however, serve as poor guidelines for

establishing the performance limits of these types of actuators.

A better approach would be to investigate the maximum achievable work density of

representative designs for each physical principle. We write the work density as a

338



function of geometrical design variables limited by a set of constraints (e.g. physical

properties and/or physical thresholds), which must all be satisfied.

Using optimization algorithms, we compared electrostatic (ES), electromagnetic

(EM) and piezoelectric (PE) actuators. Highest work densities are found in EM and

PE actuators. Even so, in the case of ES and PE actuators, there is an inevitable need

for amplification mechanisms to reach the desired stroke. On this scale this leads to

the use of flexures: a lot of the available work is lost in the movement of these

flexures, resulting in poor dynamic behaviour. Overall, in terms of dynamic

behaviour, complexity of fabrication and maximum work density, EM actuators show

to be the most suitable actuation scheme in the meso-domain.

3 Implementation

The major limitation of Lorentz actuators is their thermal weakness [3]. To address

this, as well as dealing with practicality and fabrication reasons, different

configurations were devised and simulated for implementation in a 3DOF stage. The

three free degrees of freedom needed are z, θx and θy. As we use a planar base plate,

an out-of-plane actuator is desired (Figure 1A).

Figure 1 – A: Simple 3DOF stage with the corresponding coordinate system. B:

Semi-assembled CAD model of the final design.

The final design (Figure 1B) consists of a cylindrical Nd-magnet ( 1.5 mm, h 0.5

mm, Br 1.38 T) with on top an iron yoke ( 1.5 mm, h 0.25 mm) to compress the

magnetic field and guide it outwards. Around the yoke, a flat single-turn copper coil

is located (Figure 2A), embodying the most magnetic flux in its initial position ( i

1.9 mm).

As the mass of the magnet and yoke is sufficiently small, we find that it is more

convenient to make the coil the stator, meaning it can be fabricated on a fixed

B A

339


substrate. This, together with the low resistance of the single-turn coil, addresses the

thermal limitation mentioned above, as we can now easily add cooling fins to the coil.

Furthermore, using PCB fabrication methods, we can easily manufacture a first

prototype with high enough precision. Using this method however, we are limited in

the height of the coil for the first prototype. For typical thicknesses of 35 - 70 μm, the

motor-constant is shown in Figure 2B. An optimum is found at a width of 0.4 mm.

For the proposed system, according to FEM simulations in COMSOL Multiphysics

4.2a, a 1 A current at 293.15 K, leads to a temperature increase of about 1 K.

Furthermore, temperature gradients are almost absent. This is confirmed by an

infrared measurement (Figure 3A): when subjected to a current of 0.8 A, the coils

remain at room temperature. Moreover, when a 1 A current is applied a current

density of 108 A/m2 is to be expected. When properly guided, this results in a stroke

of over 200 μm with an almost constant force of 5 mN. So not only is the motor-

constant almost independent on the stroke (and temperature), but this also means we

can use relatively simple amplifiers/circuit boards for actuation.

The yokes are cut from a 0.25 mm soft iron plate by wire-EDM. The 3DOF stage is

fabricated by exposing a 0.5 mm thick amorphous silica wafer to femtosecond laser

pulses. This induces structural modifications at the laser focal spot, due to nonlinear

absorption. Due to preferential etching, the modified regions are etched away in a hy-

drofluoric (2.5%) etching agent, after which we are left with the desired stage. Lay-

ins for the magnets ensure proper individual axial alignment (Figure 3B).

Figure 2 – A: 2D rotationally symmetric simulation of the final design. B: Motor

constant for the suggested system as a function of width for two different coil

thicknesses (35 and 70 μm).

B A

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Figure 3 – A: Infrared image of the system. When the coils and a similar resistor (0.5

Ω) are subjected to I = 0.8 A, the coils stay at room temperature, while the resistor’s

temperature increases significantly. B: Assembled system with the magnets glued

into the lay-ins of the 3D machined, transparent silica 3DOF stage.

4 Outlook

Future work includes the testing of the final prototype. Typically, in 3DOF tactile

stages, strain gauges are used for measuring the angular deflection of the stage [4].

However, using an autocollimator, we will optically investigate the angular

deflection that the actuators can generate as a function of operating frequency. By

characterizing the stage (force-displacement relation) beforehand, we have a

reference for the force delivered by the actuators.

References:

[1] Bell, D.J. et al., MEMS actuators and sensors: observations on their

performance and selection for purpose. Journal of Micromechanics and

Microengineering, 2005, 15(7): p. 153-164.

[2] Boustheen, A. et al., Active microvalves for micro-fluidic networks in plastics –

Selecting suitable actuation schemes. Microfluidics and Nanofluidics, 2011.

11(6): p. 663-673.

[3] Toma, A., Actuators enabling high-precision stages within semiconductor

equipment. Mikroniek, 2013, 53(1): p. 15-20.

[4] Gannen XP, XPRESS Precision Engineering, http://www.xpresspe.com/

B A

341

http://www.xpresspe.com/

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