Sensors and Actuators A 113 (2004) 8693

A microturbine for electric power generationJan Peirs, Dominiek Reynaerts, Filip Verplaetsen

Department of Mechanical Engineering, Katholieke Universiteit Leuven, Celestijnenlaan 300B,Leuven 3001, Belgium

Received 15 September 2002; received in revised form 25 July 2003; accepted 11 January 2004

Available online 25 February 2004

Abstract

A single-stage axial microturbine has been developed with a rotor diameter of 10 mm. This turbine is a first step in the developmentof a microgenerator that produces electrical energy from fuel. The turbine is made of stainless steel using die-sinking electro-dischargemachining. It has been tested to speeds up to 160,000 rpm and generates a maximum mechanical power of 28 W with an efficiency of18%. When coupled to a small generator, it generates 16 W of electrical power, which corresponds to an efficiency for the total systemof 10.5%. 2003 Elsevier B.V. All rights reserved.

Keywords: Microturbine; Microgenerator; Power MEMS; EDM

1. Introduction

Most portable devices use batteries for their power sup-ply. Current Li-ion batteries have energy densities up to0.5 MJ/kg but still offer limited autonomy to for instancelaptops and cellular phones. Charging times also pose prob-lems. Fuel, on the other hand, offers a much higher energydensity of about 45 MJ/kg, and the reservoir can easily berefilled. Therefore, several groups are working on the devel-opment of micro power generators based on fuel cells [13],thermo-electric devices [4,5], Stirling engines [6,7], recip-rocating internal combustion engines [8,9], Wankel motors[10], and gas turbines [1114].

Specific about the microturbine presented in this paper isthat it is an axial turbine produced with electro-dischargemachining (EDM). The microturbine developed at MIT [11]is a radial turbine with a rotor diameter of 4 mm, producedlithographically in Si or SiC. The microturbine developedat Stanford [12] is an axialradial turbine with a rotor di-ameter of 12 mm. The silicon nitride rotor is produced by agel-casting technique using a wax mould. Teams at TohokuUniversity [13] and the University of Tokyo [14] use as wellradial as axial-radial designs.

Corresponding author. Tel.: +32-16-322640; fax: +32-16-322987.E-mail address: Dominiek.Reynaerts@mech.kuleuven.ac.be(D. Reynaerts).

2. Microturbines-scale effects

The reduction of scale has several effects on the per-formance and construction of the turbine, and on the fuelchoice.

2.1. Increasing power density with miniaturisation

Dimensional analysis shows that the power P generatedby a turbomachine is proportional to the density of thegas, the fifth power of the diameter D, and the third powerof the rotational speed n:

P D5n3 (1)The power per unit volume (V D3) is thus:P

V D2n3 (2)

For a known pressure ratio and constant inlet conditions, thespeed of the fluidum at the exit of the nozzles is a constant,independent of the size of the nozzles. Therefore, the cir-cumferential speed of the turbine is constant, independentof the turbine size. This means that at optimal working con-ditions, size and rotational speed are inversely proportional:

D n = constant (3)The power density is thus inversely proportional to size:

0924-4247/$ see front matter 2003 Elsevier B.V. All rights reserved.doi:10.1016/j.sna.2004.01.003

J. Peirs et al. / Sensors and Actuators A 113 (2004) 8693 87

P

V 1

D(4)

The power density of turbomachines increases thus withminiaturisation. By replacing one turbomachine by k smallermachines with each one kth of the power, the total volumeand mass can be reduced by a factor k. This mass reductionis advantageous for aeronautical and space applications. Re-placing one large turbomachine by several smaller ones alsoimproves the reliability, especially when a few redundantdevices are added [12].

However, the above scale laws should be interpreted withcare as they are strictly speaking only valid for constantReynolds numbers. As will be shown in Section 2.3, theReynolds number decreases with miniaturisation, which hasa negative influence on the power density.

The power density of miniature turbomachines is also lim-ited for technological reasons. Small turbomachines cannotbe made with the same relative accuracy and detail as largeones, so the performance will be worse than predicted bythe scale laws.

2.2. High rotational speeds

The rotational speed is inversely proportional to the di-ameter. For a turbine diameter of 10 mm the rotational speedcorresponding to sonic flow at the outlet of the nozzles isalready 325,000 rpm. For a turbine of 5 mm diameter the ro-tational speed rises to 650,000 rpm. This is clearly beyondthe limit of ball bearings such that air or fluid bearings arerequired [15,16].

The speed is also limited by the bursting speed of therotor. This bursting speed (in terms of circumferential speed)is a constant for a certain rotor geometry and material, andthus independent of size. The burst limit corresponds thusto a rotational speed inversely proportional to size. Thus,the bursting speed and the speed resulting from the pressureratio follow the same scale law.

2.3. Low Reynolds number

The Reynolds number Re characterises the flow and isdefined as:

Re = uL, (5)

with u a characteristic speed, L a characteristic dimension ofthe flow channels, and the kinematic viscosity. The speedcan be considered independent of size as it depends onlyon the pressure ratio. The Reynolds number is thus propor-tional to size and, therefore, decreases with miniaturisation.For small turbines the flow will be less turbulent and morelaminar. This means that the viscous friction losses will behigher and that mixing of the fuelair mixture will be slower,both having a negative impact on efficiency and powerdensity.

2.4. Fast start-up and stop

The small inertia of the rotor allows start-up and stopof the turbine within a fraction of a second. This allowspower regulation using pulsed operation [17]. In that case,the turbine can operate at its optimal speed and generate afixed voltage when coupled to a generator.

2.5. Increased heat transfer

The increasing surface-to-volume ratio results in higherheat transfer. The higher thermal losses have a negative effecton the efficiency of the turbine, and may even cause flameextinction. At very small sizes, the heat generated by thecombustion minus the heat loss is not longer sufficient toignite the mixture. Another effect is that thermal insulationbetween the hot parts and the cold parts becomes more andmore a problem.

2.6. Shorter residence time

The residence time of the fuelair mixture is proportionalto the size of the turbine. This means that the time for mix-ing and combustion decreases for smaller turbines. Whena conventional turbine would be made 500 times smaller,the residence time would be reduced to the characteristickinetic reaction time of the fuelair mixture (0.010.1 ms)[18]. Therefore, the relative size of the combustion cham-bers should be increased and fuels with shorter combustiontime and shorter combustion delay should be used.

In current combustion chambers, mixing takes a large partof the residence time. Therefore, it would be an improve-ment to pre-mix air and fuel before they enter the com-bustion chamber. A disadvantage of pre-mixing is that thefuel-rich and stable primary zone in the combustion cham-ber disappears. Stable combustion of the lean mixture canbe obtained by the use of hydrogen as fuel and by the use ofcatalysts.

Hydrogen is an ideal fuel in many aspects. Comparedto hydrocarbon fuels, hydrogen has a higher mass-specificcombustion energy, a higher evaporation speed, a higher dif-fusion speed, a shorter reaction time, a considerably higherflame propagation speed, wider ignition limits, a lower ig-nition energy, and lower radiation losses. The wide ignitionlimits remove the need for a relatively rich primary combus-tion zone.

3. Turbine design

In a first phase of the project, the problem has been scaleddown to a turbine powered by compressed air. Compres-sor, combustion chamber, and generator have been left outand will be addressed in a later phase. The microturbine isa single-stage axial impulse turbine (Laval turbine). Expan-sion of the gas takes place in the stationary nozzles and not

88 J. Peirs et al. / Sensors and Actuators A 113 (2004) 8693

1 2 3 4 5 4 6 7 8

Fig. 1. Microturbine design.

between the rotor blades. This type of turbine has been cho-sen because of its simple construction.

Fig. 1 shows an exploded view and an assembly of themicroturbine design. The compressed air enters via a stan-dard pneumatic connector (1) and expands over the station-ary nozzles (3) where it is deflected in a direction tangentialto the turbine rotor (5). After the air has passed the rotorblades, it leaves the device through the openings in the outletdisc (6). Screwing the pneumatic connector in the housing(8) presses the stationary nozzle disc against a shoulder inthe housing. The rotor blades, wheel and axis are one mono-lithic part. The rotor is supported by two ball bearings (4),one mounted in the stationary nozzle disc and one mountedin the outlet disc. The outlet disc is locked in the housingby a circlip (7).

The diameter of the turbine rotor is 10 mm. The housinghas a diameter of 15 mm and is 25 mm long. All parts, ex-cept the pneumatic connector and the circlip, are made ofstainless steel.

The nozzles are designed for subsonic flow and, therefore,have a converging cross section. Sonic speed is reached for arelative supply pressure of 1 bar. The exit losses (remainingkinetic energy in the exhaust) are minimal when the turbineis designed for a u/c1 ratio of 0.5, with u the circumferentialspeed and c1 the absolute speed at the nozzle exit. At 1 bar,c1 reaches sonic speed resulting in an optimal turbine speedof 420,000 rpm. As this is too high for the bearings, theturbine has been designed for a u/c1 ratio of 0.25, and isoperated below its optimal speed of 210,000 rpm.

4. Turbine production

The different parts of the turbine are produced by turningand EDM. The nozzle disc and rotor are the most complexparts. In a first step, their cylindrical surfaces are machinedon a lathe. In a second step, the nozzles and blades arecreated by die-sinking EDM as illustrated for the rotor inFig. 2. The rotor is clamped in a rotary head which is indexedwith steps of 30. A prismatic copper electrode with a crosssection having the shape of the air channels between theblades is sunk into the turbine wheel by EDM. The electrode

Fig. 2. Machining of the rotor blades by EDM.

Fig. 3. Subassembly of nozzle disc, turbine rotor, and bearings. The rotorhas a diameter of 10 mm.

is produced by wire-EDM. Fig. 3 shows a subassembly ofnozzle disc, rotor, and bearings.

5. Generator

The turbine has been coupled to a small brushless dcmotor that is used as a three-phase generator (see Fig. 4).The motor (Faulhaber, type 1628 T024B K312) has aboutthe same size as the turbine: 16 mm in diameter and 28 mmlong. Turbine and generator are coupled to each other byan elastic tube that serves as a flexible coupling.

6. Mechanical output

Torque and power of the turbine have been tested up to aspeed of 100,000 rpm. For this purpose, a 30 mm diameter

J. Peirs et al. / Sensors and Actuators A 113 (2004) 8693 89

Fig. 4. Turbine coupled to the generator.

brass wheel has been fixed to the turbine axis, as shownin Fig. 5. An optical sensor measures the rotation of thewheel in a contactless way: two vanes on the wheel inter-rupt the optical path of a photosensor. The turbine is testedby switching on the pressure and accelerating the turbine to100,000 rpm. The torque is then derived from the accelera-tion and the moment of inertia of the wheel and turbine rotor.As the turbine passes through the whole speed range, accel-eration, torque and power are known as a function of speed.

When the turbine is rotating at full speed, the pressure isswitched off and a new measurement is done while the tur-bine slows down. This gives the friction torque as a functionof speed. Friction mainly occurs between the wheel withvanes and the surrounding air. The friction torque and powerare added to the results of the acceleration test to obtain thetotal torque and power of the turbine.

Figs. 6 and 7 show torque and mechanical power as afunction of speed for different supply pressures up to 1 bar.The maximum torque and power are, respectively, 3.7 Nmmand 28 W. The dashed lines represent the friction losses de-termined with the deceleration test.

At 1 bar, the turbine consumes 8 Nm3/h of compressedair, which corresponds to a power consumption of 152 Wwhen assuming an ideal isentropic expansion. This meansthat the mechanical efficiency of the turbine lies around 18%.Fig. 8 shows the turbine efficiency as a function of speedfor different supply pressures.

The dips in the characteristics at high speed are causedby the measurement method as they always occur at the

Turbine

Inertia wheel

Vane

Photosensor

Fig. 5. Set-up to measure the mechanical output of the turbine. The outputtorque is derived from the acceleration of the inertia wheel.

0 20 40 60 80 100 120

x 103

Torque(Nmm)

Speed (rpm)

0

0.5

1

1.5

2

2.5

3

3.5

1 bar

0.8 bar

0.7 bar

Friction t

0.6 bar

0.5 bar

0.2 bar

orque

Fig. 6. Torque generated by the turbine as a function of speed and supplypressure.

Mechanicalpower(W)

0 20 40 60 80 100 120

x 103

Speed (rpm)

0

5

10

15

20

25

30

1 bar

0.8 bar

0.7 bar

0.6 bar

0.5 bar

0.2 bar

Friction loss

Fig. 7. Mechanical power of the turbine as a function of speed and supplypressure.

Turbineefficiency(%)

0 20 40 60 80 100 120

x 103

Speed (rpm)

0

2

4

6

8

10

12

14

16

18

20

1 bar

0.8 bar

0.7 bar0.6 bar0.5 bar

Fig. 8. Efficiency of the turbine (compressed air to mechanical power).

90 J. Peirs et al. / Sensors and Actuators A 113 (2004) 8693Electricalpower(W)

x 103

Speed (rpm)

0 20 40 60 80 100 120 140 1600

2

4

6

8

10

12

14

16

18

1 bar

0.9 bar

0.8 bar

0.7 bar

0.6 bar

0.5 bar

0.4 bar

0.3 bar0.2 bar

Fig. 9. Electrical power generated by the total system (turbine plusgenerator).

maximal speed, even for different loads and pressures. Inreality, power and efficiency increase further with speed toreach their maxima theoretically at 210,000 rpm (for 1 bar).These speeds can be reached using a smaller load.

7. Electrical output

To measure the electrical power output of the system,the generator is connected to a variable three-phase loadconsisting of three potentiometers (range: 2 k, 10 turns).In contrast with the mechanical tests, the electrical tests areperformed at constant speed. The speed of the turbine, whichis measured from the frequency of the generator voltage, iscontrolled by varying the load. Fig. 9 shows the electricalpower measured for different supply pressures and speeds.At a pressure of 1 bar, the maximal electrical power is 16 Wand is reached at a speed of 100,000 rpm. Measurementsshow that the air flow and input power depend only on thesupply pressure and not on speed or load. Therefore, theinput power is the same as in the mechanical test at 1 bar,i.e. 152 W. Fig. 10 shows the total efficiency (compressedair to electricity) as a function of speed and for differentsupply pressures. The maximal total efficiency is 10.5% andis reached at a speed of 100,000 rpm.

8. Sankey diagram

The energy flow and the different losses are illustratedin the Sankey diagram shown in Fig. 11. The diagramis generated for a supply pressure of 1 bar and a speedof 100,000 rpm. This corresponds to the working pointat which the maximal electrical power and maximal totalefficiency are reached. Input power, mechanical power,electrical power and the combination of ventilation losses(6) and bearing friction (7) are measured values. This last

Totalefficiency(%)

x 103

Speed (rpm)

0 20 40 60 80 100 120 140 1600

2

4

6

8

10

12

1 bar

0.9 bar0.8 bar0.7 bar

0.6 bar

0.5 bar

0.4 bar

0.3 bar

0.2 bar

Fig. 10. Total efficiency (compressed air to electricity).

value (6 + 7) is obtained with a deceleration test of theturbine without generator and without external load. Theloss associated with the leak flow around the turbine wheel(2) and the exit losses (8) are calculated from the knownair speeds. The expansion losses (1), incidence losses (4)and blade profile losses (5) are calculated using friction andloss coefficients known from large turbines and may be less

Expansion losses15 W - 9.8 %

Leak flow around rotor4 W - 2.6 %Obstruction losses1 W - 0.7 %Incidence losses2 W - 1.3 %

Blade profile losses48 W - 31.6 %

Ventilation losses +bearing friction2 W - 1.3 %

Exit losses52 W - 34.2 %

Losses in coupling2 W - 1.3 %

Generator losses10 W - 6.6 %

Mechanicalpower

28 W - 18.4 %

Electrical power16 W - 10.5 %

Input power152 W - 100 %

Measured

Legend

Difference fromother valuesCalculated

Fig. 11. Sankey diagram for a supply pressure of 1 bar and a speed of100,000 rpm.

J. Peirs et al. / Sensors and Actuators A 113 (2004) 8693 91

accurate. The generator losses (10) are derived from themanufacturers data sheets. The obstruction losses (3) andthe losses in the coupling (9) are derived as the differencebetween the calculated and measured values.

The major losses are the blade profile losses and the exitlosses. The large blade profile losses can be explained bythe increased friction in miniature systems (small channelsand low Reynolds numbers). The high exit losses can beexplained by the low u/c1 ratio (0.25 instead of 0.5 in theoptimal case). Additionally, the turbine operates below itsoptimal speed because the ball bearings limit the speed. Bothfactors result in higher air speeds at the turbine exit, andthus higher exit losses.

9. Power density

Table 1 gives the masses of the different parts. The tur-bine housing and pneumatic connector are not optimised to-wards mass and are responsible for about 86% of the weight.Therefore, the mass of the turbine can be substantially re-duced by optimising these parts.

The mechanical power density of the turbine, defined asthe mechanical power output of the turbine (28 W) dividedby the mass of the turbine (36 g), is about 780 W/kg. As acomparison, Fig. 12 shows the power density of commercialgas turbines for helicopters, tanks, ships and power gener-ators from General Electric, Rolls-Royce, and Pratt & Wit-ney. Most of these turbines have a power density between 4and 10 kW/kg. The tank turbine has a lower power densityof 1 kW/kg. The power density of the current microturbineis thus 510 times lower than the power density of large tur-bines. This figure can be improved by optimising the massof the connector and housing, but on the other hand, a com-pressor and combustion chamber have to be added to obtaina turbine comparable to the ones mentioned in Fig. 12. Acomparison can also be made with a silicon air turbine witha rotor diameter of 4.2 mm developed at MIT [15]. It gen-erates 5 W of power and achieves a power density of morethan 4 MW/m3 (about 2 kW/kg), about 2.5 times more thanthe turbine presented in this paper.

Table 1Masses of the different parts

Part Mass (g)Turbine 36Pneumatic connector (1) 15.8Ring (2) 0.77Nozzle disc (3) 1.78Small bearing (4) 0.03Large bearing (4) 0.07Rotor (5) 1.63Outlet disc (6) 0.35Circlip (7) 0.27Housing (8) 15Generator 30

Total (turbine + generator) 66

1

10

100

10 100 1000 10000 100000

Mass (kg)

Power/mass(kW/kg)

Helicopter

Tank

Ship

Power system L-1

M-1/3

Fig. 12. Power density of commercial gas turbines. Data from GeneralElectric, Rolls-Royce, and Pratt & Witney. The line represents the evo-lution of the power density as predicted by the scale laws.

The line shown in Fig. 12 shows the evolution of thepower density as predicted by the scale laws derivedabove. It is clear that the data does not correspond to thisscale law. Extrapolation predicts even a power density of100300 kW/kg for the current microturbine, more than twoorders of magnitude above the measured values. As men-tioned in Section 2.1, this can be explained by technologicallimitations and increased friction losses.

The electrical power density of the microturbine, definedas the electrical power produced by the generator (16 W)divided by the combined masses of turbine and generator(66 g), is about 240 W/kg.

10. Conclusion

A 10 mm diameter axial microturbine with generator hasbeen developed and successfully tested to speeds up to160,000 rpm. It generates a maximum mechanical power of28 W with an efficiency of 18%. Power and efficiency aremainly limited by the maximal speed of the ball bearings.The main losses are the blade profile losses (32%) and theexit losses (34%). Higher speeds can considerably reducethe exit losses and therefore increase efficiency and power.Currently, the power density is 780 W/kg, about 510 timeslower than for large turbines. However, higher speeds andoptimisation of the housing can considerably increase thisfigure. When coupled to a small generator, the system gen-erates 16 W of electrical power, corresponding to a totalefficiency of 10.5%.

11. Future work

The first goal is to increase the efficiency of the tur-bine, mainly by decreasing the exit losses. Therefore,

92 J. Peirs et al. / Sensors and Actuators A 113 (2004) 8693

the allowable speed of the turbine will be increased to200,000300,000 rpm by using special high-speed ballbearings. Later on, air or fluid bearings can be intro-duced to attain even higher speeds. Another possible ap-proach is to decrease the speed by using a multiple-stagedesign.

A compressor and a combustion chamber will be addedto finally come to a microgenerator running on fuel. Thecompressor is currently under development.

Acknowledgements

This research is sponsored by the Belgian programmeon Interuniversity Poles of Attraction (IAP5/AMS) initiatedby the Belgian State, Prime Ministers Office, Science Pol-icy Programming. The authors assume the scientific respon-sibility of this paper. The authors wish to thank MichaelPoesen and Pieterjan Renier for their contribution to thiswork.

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Biographies

Jan Peirs Graduated as mechanical engineer (K.U. Leuven, 1993). Hestarted his activities as a research assistant at the Division of ProductionEngineering, Machine Design and Automation (PMA) of K.U. Leuvenin 1993. He received a PhD degree in mechanical engineering in 2001from K.U. Leuven. Currently, he is working as a Postdoctoral researcherat PMA. His research interests include the design of micro-actuators,medical microsystems, micro-powergenerators, and micromechanicalsystems in general.

Dominiek Reynaerts Graduated as mechanical engineer (K.U. Leuven,1986). He started his activities as a research assistant at the Divisionof Production Engineering, Machine Design, and Automation of K.U.Leuven in 1986. Within the framework of the Erasmus student exchangeprogram, he stayed as a PhD student at the Scuola Superiore S. Annain Pisa in 1990. In 1993, he became a research manager of the divisionPMA. He received a PhD degree in mechanical engineering in 1995 fromK.U. Leuven. Currently, he is associate professor at K.U. Leuven. Hiscurrent research interests include design and control of multi-fingeredrobot grippers, shape memory alloy actuators, precision mechanics, andmicromechanical systems.

J. Peirs et al. / Sensors and Actuators A 113 (2004) 8693 93

Filip Verplaetsen received his Engineering degree from the Univer-siteit Gent, Ghent, Belgium, in 1992 and the Diplome en Adminis-tration des Entreprises from the Universit Catholique de Louvain,Louvain-la-Neuve, Belgium, in 1993. He joined the Katholieke Univer-siteit Leuven , Louvain, Belgium in 1994 as a research assistant andreceived the PhD degree in mechanical engineering in 1999. From 1999

till 2002, he worked as a Postdoctoral researcher of the Fund for Scien-tific Research, Flanders (F.W.OVlaanderen) at the same university andbecame assistant professor in 2002. His research focuses on industrialsafety, explosion safety, heat transfer enhancement techniques, design ofthermal systems and turbomachinery.

A microturbine for electric power generationIntroductionMicroturbines-scale effectsIncreasing power density with miniaturisationHigh rotational speedsLow Reynolds numberFast start-up and stopIncreased heat transferShorter residence time

Turbine designTurbine productionGeneratorMechanical outputElectrical outputSankey diagramPower densityConclusionFuture workAcknowledgementsReferences