Ansi Agma6010 f97

ANi!U/AGMA 6010- F97 (Revision of

ANSI/AGMA 601 O-E88)

AMERICAN NATIONAL STANDARD

t -

Standard for Spup; Helical, Herringbone and Bevel Enclosed Drives

AGMA STANDARD = -= Reproduced By GLOBAL

= - ENGINEERING DOCUMENTS -- B g Wth The Permission Of AGMA

c? Under Royalty Agreement

-

American Standard for Spur, Helical, Herringbone and Bevel Enclosed Drives ANSI/AGMA 601 O-F97

National [R evision of ANWAGMA 601 O-E881

Standard Approval of an American National Standard requires verification by ANSI that the requirements for due process, consensus and other criteria for approval have been met by the standards developer. Consensus is established when, in the judgment of the ANSI Board of Standards Review, substantial agreement has been reached by directly and materially affected interests. Substantial agreement means much more than a simple majority, but not necessarily unanimity. Consensus requires that all views and objections be considered, and that a concerted effort be made toward their resolution. The use of American National Standards is completely voluntary; their existence does not in any respect preclude anyone, whether he has approved the standards or not, from manufacturing, marketing, purchasing or using products, processes or procedures not conforming to the standards. The American National Standards Institute does not develop standards and will in no circumstances give an interpretation of any American National Standard. Moreover, no person shall have the right or authority to issue an interpretation of an American National Standard in the name of the American National Standards Institute. Requests for interpretation of this standard should be addressed to the American Gear Manufacturers Association. CAUTION NOTICE: AGMA technical publications are subject to constant improvement, revision or withdrawal as dictated by experience. Any person who refers to any AGMA technical publication should be sure that the publication is the latest available from the As- sociation on the subject matter. Fables or other self-supporting sections may be quoted or extracted. Credit lines should read: Extracted from ANSI/AGMA 6010-F97, Standard for Spur, Helical, Herringbone and Bevel Enclosed Drives, with the permission of the publisher, the American Gear Manufacturers Association, 1500 King Street, Suite 201, Alexandria, Virginia 22314.1

Approved October 2,1997

ABSTRACT This standard includes design, rating, lubrication, testing and selection information for spur, helical, herringbone and bevel gears when using enclosed speed reducers or increasers. Units covered include those with a pitch line velocity below 7000 feet per minute or rotational speeds no greater than 4500 rpm.

Published by

American Gear Manufacturers Association 1500 King Street, Suite 201, Alexandria, Virginia 22314

Copyright 0 1997 by American Gear Manufacturers Association All rights reserved.

No part of this publication may be reproduced in any form, in an electronic retrieval system or otherwise, without prior written permission of the publisher.

Printed in the United States of America

ISBN: l-55589-890-1

ii

AMERICAN NATIONAL STANDARD ANSIIAGMA 601 O-F97

Contents Page

Foreword ............................................................... iv 1 Scope .............................................................. 1 2 Symbols, terminology and definitions .................................... 1 3 Application and design considerations .................................... 4 4 Unitrating ........................................................... 5 5 Rating criteria ....................................................... 6 6 Thermal power rating ................................................. 8 7 Componentdesign .................................................. 15 8 Service factors ...................................................... 18 9 Lubrication and lubricants ............................................ 19 10 Assembly and rotation ............................................... 22 11 Ratios ............................................................. 24 12 Marking ............................................................ 24 13 Storage ............................................................ 24 14 Installation ......................................................... 25

Tables

1 Symbols used in equations ............................................ 2 2 Bearing coefficient of friction, fb ....................................... 11 3 Lubricant factor, C1, at 200°F sump temperature ......................... 11 4 Heat transfer coefficient, k, for gear drives without auxiliary cooling ......... 14 5 Heat transfer coefficient, k, for gear drives with fan cooling ................ 14 6 Ambient temperature modifier, B,f ..................................... 14 7 Ambient air velocity modifier, & ....................................... 14 8 Attitude modifier, BA ................................................. 15 9 Maximum allowable oil sump temperature modifier, & .................... 15 10 Operation time modifier, 80 ........................................... 15 11 Nominalratios ...................................................... 24

Figures

1 Seal friction torque .................................................. 12 2 Shaft rotation ....................................................... 22 3 Parallel shaft spur, helical and herringbone gear drives, single or multiple

stage .............................................................. 22 4 Horizontal bevel gear drives, single stage; horizontal bevel-helical drives,

multiple stage. ...................................................... 23 5 Vertical bevel gear drives, single stage; vertical bevel-helical drives, multiple

stage .............................................................. 23

Annexes

A Service factors ...................................................... 29 B Keys and keyways for shaft extensions ................................. 37 C lllustrativeexamples ................................................. 39 D Test and inspection procedures ....................................... 49 E Owner responsibilities ............................................... 51 F Gear tooth mesh losses for bevel gears ................................ 53

Bibliography . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 55

. . . 111

ANSI/AGMA 601 O-F97 AMERICAN NATIONAL STANDARD

Foreword

rhe foreword, footnotes and annexes, if any, in this document are provided for informational purposes only and are not to be construed as a part of ANSI/AGMA Standard 6010-F97, Standard for Spur, Helical, Herringbone and Bevel Enclosed Drives.]

This standard presents general guidelines and practices for design, rating (including catalog rating) and lubrication of enclosed gear drives and is a revision to and supersedes ANSIIAGMA 6010-E88, Standard for Spur, Helical, Herringbone and Bevel Enclosed Drives.

This standard was revised to include the latest data available using current gear technology and includes operational experience. Also, this revision conforms to the American Gear Manufacturers Association style manual. This standard is written in conventional units while the parallel standard, ANSVAGMA 61 lo-F97, is available in hard metric.

The allowable stress numbers used in this standard are derived from ANSVAGMA 2001 -C95 and ANSVAGMA 2003-A86, and along with other rating factors, provide a rating basis for enclosed gear reducers and increasers. The rating formulas are based on many years of experience in the design and application of enclosed gear drives for industrial use. The allowable stress numbers used by this standard are adjusted by the factors .& and YN for the required number of cycles of operation, In previous versions of this standard, a CL and & value of unity has been allowed. Deviations from ANSVAGMA 2001-C95 or ANSI/AGMA 2003-A86 are not recommended unless they can be justified. The use of the stress cycle adjustment factor does not guarantee that a certain number of hours or revolutions of life will be obtained, but is a method of approximating gear life under different load and speed conditions.

The most significant changes in this standard include: information on obtaining ratings by direct reference to the empirical methods in ANSVAGMA 2001-C95 and ANSVAGMA 2003-A86; references to the necessary additional standards; a uniform selection method by specifying a nominal L1 life for the gearing of 10 000 hours; and a refinement of the thermal rating practice. The competence to design enclosed gear drives, especially the knowledge and judgment required to properly evaluate the various rating factors, comes primarily from years of experience in designing, testing, manufacturing and operating similar gear drives. The proper application of the general rating formulas for enclosed gear drives is best accomplished by those experienced in the field. There is a need for a thorough knowledge and use of the safety, service and application factors. Application factors will represent actual loadings or will be replaced by a load spectrum analysis such as M iner’s Rule.

Work was started on this draft in Decemberl991. This version was approved by the AGMA membership in June 1997. It was approved as an American National Standard on October 2, 1997.

Suggestions for improvement of this standard will be welcome. They should be sent to the American Gear Manufacturers Association, 1500 King Street, Suite 201, Alexandria, Virginia 22314.

iv

AMERICAN NATIONAL STANDARD ANSI/AGMA 601 O-F97

PERSONNEL of the AGMA Helical, Herringbone and/or Spiral Bevel Enclosed Drives and Standard Units Committee

Chairman: W.P. Crosher .................... Flender Corporation Vice Chairman: G.A. DeLange ............... Prager, Inc.

ACTIVE MEMBERS

R.L.Cragg ................................ Consultant R.G. Ferguson ............................ Amarillo Gear Company R.W. Holzman ............................. Milwaukee Gear Company, Inc. HR. Johnson, III ........................... Lufkin Industries, Inc. L. Lloyd .................................. Lufkin Industries, Inc. D.L. Mairet ................................ Consultant D. McCarthy .............................. Dorris Company D.R. McVittie .............................. Gear Engineers, Inc. A.R. Perry ................................ Dorris Company A.E. Phillips ............................... Rockwell Automation/Dodge V.Z. Rychlinski ............................ Brad Foote Gear Works, Inc. B.W. Shirley. .............................. Emerson Power Transmission Corp. R.G.Smith ................................ Philadelphia Gear Corporation F.C. Uherek ............................... Flender Corporation

ASSOCIATE MEMBERS

J.F. Alison, III ............................. Steward Machine Company, Inc. R.G. Allenby .............................. Hamilton Gear A.C. Becker ............................... Nuttall Gear Corporation K.A. Beckman ............................. Lufkin Industries, Inc. A.S. Cohen ............................... Engranes y Maquinaria Arco D. Fleischer ............................... Hamilton Gear, Inc. R.A. Geary ............................... LCI, Inc. J. Gimper ................................. Danieli United, Inc. B. Goebel ................................ The Horsburgh & Scott Company . lvers ................................... Xtek, Inc. D. King ................................... D.L. King &Associates C.E. Long ................................ Cummins Engine G. McCain ................................ Amarillo Gear Company J.R. Partridge ............................. Euro Lufkin bv M. Peculis ................................ The Horsburgh & Scott Company W.P. Pizzichil .............................. Philadelphia Gear Corporation R.K. Polen ................................ The Alliance Machine Company P.N. Salvucci .............................. IMO Industries, Inc. M.D. Schutte .............................. Lightnin E.S. Scott ................................ The Alliance Machine Company J. Simpson, Jr. ............................ Turner Uni-Drive Company L. Spiers ................................. Emerson Power Transmission I. Wilson .................................. WesTech Gear Corporation S. Yamada ................................ Sumitomo K. Yasui .................................. Seiki-Kogyosho, Ltd. (SKK)

V

ANSI/AGMA 601 O-F97

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AMERICAN NATIONAL STANDARD ”


American National S tandard -

Standard for Spur, Helical, Herringbone and Bevel Enclosed Drives

1 Scope

This standard is applicable to enclosed gear drives wherein the gear tooth designs include spur, helical, herringbone or double helical, or bevel, in single or multistage units.

I .I Limitations

The rating methods and influences identified in this standard are lim ited to enclosed drives of single and multiple stage designs where the pitch line velocities do not exceed 7000 feet per m inute and pinion speeds do not exceed 4500 revolutions per m inute.

I .2 Overlap

There is a speed and pitch line velocity overlap in the scope of this standard and thescope of ANWAGMA 6011 -G92. ANSI/AGMA6011 -G92 permits designs down to 5000 feet per m inute and pinion speeds to 3600 revolutions per m inute. In the areaof overlap in the scope, the standard used is dependent on the application or unit designed. All parties should be aware of the standard used and should agree on the use of that standard.

I.3 Intended use

This standard is not intended to assure performance of assembled gear drive systems. It is intended for use by the experienced gear designer capable of selecting reasonable values for the factors, based on his knowledge of performance of similar designs and the effects of such items as lubrication, deflection, manufacturing tolerances, metallurgy, residual stress and system dynamics. It is not intended for use by the engineering public at large.

I .4 Exceptions

This standard does not cover the design and application of epicyclic drives or gear blank design.

This standard does not cover the rating of gear drives due to wear or scoring (scuffing) of gear teeth or components.

This standard does not apply to gear drives that are covered by other specific AGMA application standards.

I.5 Annexes

The annexes are for reference only and are not a part of this standard. The annexes can be used to make a more detailed analysis of certain rating factors, and a guide to owner responsibilities.

2 Symbols, terminology and definitions

The symbols used in this standard are shown table 1.

in

NOTE: The symbols, terms and definitions contained in this document may vary from those used in other AGMA standards. Users of this standard should assure themselves that they are using these symbols and definitions in the manner indicated herein.

2.1 Definitions

The terms used, wherever applicable, conform to the following standards:

ANSI Y10.3-1968, Letter Symbols for Quantities Used in Mechanics of Solids

ANSVAGMA IO1 2-F90, Gear Nomencla@re, Definitions of Terms with Symbols

ANWAGMA 9005-D94, Industrial Gear Lubrication

2.2 Reference documents

The following standards contain provisions which, through reference in this text, constitute provisions of this American National Standard. At the time of publication, the editions indicated were valid. All standards are subject to revision, and parties to agreements based on this American National Standard are encouraged to investigate the possibil- ity of applying the most recent editions of the standards indicated below.

ANSI/AGMA 6010-F97

AGMA 299.01, Section Ill, Gear Sound Manual: Gear Noise Control

ANSI/AGMA lOlO-E95,AppearanceofGearTeeth ANSI/AGMA 6001-D97, Design and Selection of - Terminology of Wear and Failure Components for Enclosed Gear Drives

ANSI/AGMA 1012-F90, Gear Nomenclature Definitions of Terms with Symbols

ANSIJAGMA 6025-C90, Sound for Enclosed Helical, Herringbone, and Spiral Bevel Gear Drives

ANSl/AGMA 2001 -C95, Fundamental Rating Fac- tors and Calculation Methods for lnvolute Spur and Helical Gear Teeth

ANSI/AGMA 6011 -G92, Specification for High Speed Helical Gear Units

ANSI/AGMA 2003A86, Rating the Pitting Resist- ance and Bending Strength of Generated Straight Bevel, Zero1 Bevel, and Spiral Bevel Gear Teeth

ANSI/AGMA 6000-896, Specification for Measurement of Linear Vibration on Gear Units

ANSIIAGMA 9002-A86, Bores and Keyways for Flexible Couplings (Inch Series)

ANSVAGMA 9005D94, Industrial Gear ANSI/AGMA 2008-B90, Assembling Bevel Gears Lubrication

Table 1 - Symbols used in equations

First Reference Symbol Term Units used clause

A Arrangement constant -- Eq 19 6.5.1.4 A, Gear case surface area exposed to ambient air ft2 Eq 30 6.5.2 B Length through bore of bearing in Eq 24 6.5.1.5 BA Altitude correction factor -- Eq 31 6.6 BD Operation cycle correction factor -- Eq 31 6.6 B ref Ambient temperature correction factor -- Eq 31 6.6 BT Maximum allowable sump temperature correction factor - - Eq 31 6.6 Bv Ambient air velocity correction factor -- Eq31 6.6 ccl External dynamic factor (bevel) -- -- 5.1.2 cb Stress adjustment factor (bevel) -- -- 5.1.2 Cf Surface condition factor (bevel) -- -- 5.1.1 CL Life factor (bevel) -- -- 5.1.2.1 Gtl Load distribution factor (bevel) -- -- 5.1.2 C,, Mesh alignment factor -- -- 5.1.1.2 CP Lubricant absolute viscosity CP Eq 23 6.5.1.4 G Size factor (bevel) -- -- 5.1.2 CT Temperature factor (bevel) -- -- 5.1.2 G Dynamic factor (bevel) -- -- 5.1.2 Cl Lubricant factor -- Eq 13 6.5.1.2.1 DR Mean diameter of tapered roller in Eq 24 6.5.1.5 DS Shaft seal diameter in Fig 1 6.5.1.3 4 Bearing bore in Eq 11 6.5.1 .l 4?l Mean diameter of a tapered roller bearing in Eq 24 6.5.1.5 d, Bearing outside diameter in Eq 11 6.5.1.1 49 Operating pitch diameter of gear or pinion in Eq 19 6.5.1.4 EP Electric power consumed hp Eq 29 6.5.1.6 e Tapered bearing calculation factor -- Eq 25 6.5.1.5 em Electric motor efficiency % Eq 29 6.5.1.6 ep Oil pump efficiency % Eq 28 6.5.1.6

(continued)


Table 1 (continued)

Symbo F

Ft

f6 fm H, Ht K

Kl Iii K, K3 &7 & &a

lvsf KT

&

K5 k

M

% NC

NP n

?P nb

t?4 P ac

Pat

PB

PL

P M P me 87lJl PN

PIl PP

P P m

PPS

PQ ps

PT

bhf?Z

pv

P W B

Term Face width in contact with mating element Total face of gear or pinion Bearing coefficient of friction Mesh coefficient of friction Siding ratio at start of approach Sliding ratio at end of recess Contact load factor for pitting resistance External dynamic factor (bevel) Stress cycle factor (bevel) Load distribution factor (bevel) Overload factor Reliability factor Size factor (bevel) Shrink adjustment factor to compensate for less than 100 percent engagement Service factor Temperature factor Dynamic factor Tapered bearing dynamic load ratio factor Heat transfer coefficient Mesh mechanical advantage Sear ratio Number of gear teeth Number of pinion teeth Shaft speed ‘inion speed 3earing shaft speed Application power of the enclosed drive 4llowable transmitted power for pitting resistance Uowable transmitted power for bending strength 3earing power losses -oad-dependent power losses Sear mesh power losses Vlinimum component power rating Aean normal diametral pitch Jon-load dependent power losses Jormal diametral pitch otal oil pump power required (all pumps) rlotor driven oil pump losses ihaft driven oil pump losses ieat dissipated Iii seal power losses basic thermal rating of the drive application thermal rating feat generated (total power loss) learing combined windage and churning power losses

Units in in

-- -- -- --

lb/in* -- -- -- -- -- -- --

-- -- -- --

hp/(fi*“f) -- -- -- --

rpm rw wm hp f-v hp hp hp hp hp

in-l hp

in-l hp hp hp hp hp hp hp hp hp

First used

Eq 14 Eq 19 Table 2 Eq 12 Eq 15 Eq 15 Eq 13 -- -- -- -- -- -- Eq 32

Eq 1 -- -- Eq 26 Table 4 Eq 12 Eq 16 Eq 14 Eq 14 Eq 18 Eq 12 Eq 10 Eq 1 -- --

Eq 8 Eq3 Eq 8 Eq 1 zq 21 :q 3 fq 19 Eq 9 :q 27 :q 27 !q 2 Eq 9 fq 7 :q 31 fq 2 fq 9

Reference clause

6.5.1.2.1 6.5.1.4 6.51 .l 6.5.1.2.1 6.5.1.2.1 6.5.1.2.1 6.5.1.2 5.2.2 5.2.2 5.1.1 5.1 .l 5.1.1 5.2.2 7.4.2

3.1 5.1.1 5.1.1 6.5.1.5 6.5.2 6.5.1.2.1 6.5.1.2.1 6.5.1.2.1 6.5.1.2.1 6.5.1.3 6.5.1.2.1 6.5.1.1 4.3 5.1 5.2 6.5.1 6.5 6.5.1 4.3 5.5.1.4 5.5 5.5.1.4 3.5.1 3.5.1.6 3.5.1.6 5.5 3.5.1 3.5 3.6 3.5 1.5.1

(continued)

3

ANSVAGYA 6010-F97 AMERICAN NATIONAL STANDARD

Table 1 (concluded) First Reference

Symbol Term Units used clause pw Gear combined windage and churning power loss hp Eq 9 6.5.1 P Operating oil pressure lb/in2 Eq 28 6.5.1.6 Q Oil flow . wm Eq 28 6.5.1.6 Qv Transmission accuracy level number -- -- 5.1.1.1 % Gear outside radius in Eq 16 6.5.1.2.1 RW Gear operating pitch radius in Eq16 6.5.1.2.1 rrn Mean reference radius in Eq 21 6.5.1.4 r, Pinion outside radius in Eq 17 6.5.1.2.1 rW Pinion operating pitch radius in Eq 14 6.5.1.2.1 S at Allowable contact stress number a- -- 5.1.2.2 S to Calculated bending stress number excluding dynamic - - - - 5.1.2.2

effects SF Safety factor for bending strength -- -- 5.2.1 SH Safety factor for pitting resistance -- -- 5.1.1 say Allowable yield strength number lb/in2 - - 4.4 SC Calculated key compressive stress lb/in2 Eq 32 7.4.2 s,, Adjusted compressive stress lb/in2 Eq 32 7.4.2 %k Calculated key shear stress lb/in2 Eq 33 7.4.2 %O Adjusted shear stress lb/in2 Eq 33 7.4.2 TD Design pinion torque lb in - - 5.1.2.3 TP Operating pinion torque lb in - - 5.1.2.3

Tp Torque on the pinion lb in Eq 12 6.5.1.2.1 Tb Rolling bearing friction torque lb in Eq 10 6.5.1.1 ?-f Allowable percentage of torque capacity obtained due to % Eq 32 7.4.2

interference fit z Oil seal torque lb in Eq 18 6.5.1.3 V Pitch line velocity ft/m in Eq13 6.5.1.2.1 W Equivalent radial bearing load lb Eq 11 6.5.1 .l yhi Stress cycle factor for bending strength -- -- 5.2.1 ZN Stress cycle factor for pitting resistance -- -- 5.1.1

it Cup angle of the tapered roller bearing degrees Eq 24 6.5.1.5 Operating transverse pressure angle degrees Eq 15 6.5.1.2.1

9 Mean spiral angle degrees Eq 21 6.5.1.4 WY Operating helix angle at operating pitch diameter degrees Eq 12 6.5.1.2.1 AT Temperature differential “F Eq 30 6.5.2

11 Overall unit efficiency % Eq 6 6.5

3 Application and design considerations Units rated to this standard can accommodate the following peak load conditions:

Users of this standard are expected to have had experience in the field of gearing and mechanical drive systems.

- Each peak shall not exceed 200 percent of the unit rating (service factor, K$ = 1 .O);

3.1 Application lim itations - A lim ited number of stress cycles, typically less than 1 04.

In this standard, the unit rating is defined as the mechanical capacity of the gear unit components determined with a unity service factor.

For applications exceeding these conditions an appropriate service factor should be selected.

4

AMERICAN NATIONAL STANDARD ANSIIAGMA 6010-F97

WARNING: For momentary torques in excess of 200%, stall conditions and low number of stress cycles, the gear drive should be evaluated to assure the user of this standard that these conditions do not exceed the yield strength of any component.

Some applications may require selecting a gear drive with increased mechanical rating in order to accommodate adverse effects of environmental conditions, thermal capacity of the unit, external loading or any combination of these factors such as overhung, transverse and thrust loads.

3.2 Rating factors

The allowable stress numbers taken from ANSI/ AGMA 2001 -C95 and ANSVAGMA 2003-A86 are maximum allowed values. Some latitude based upon experience is permissible in the selection of specific factors within this standard. Less conservative values for rating factors in this standard shall not be used.

Ratings shall be as outlined in clauses 4 and 5.

3.3 Cold temperature operation

If units are to be operated below -2O”F, care must be given to select materials which have adequate impact properties at the operating temperature. Consideration should be given to:

- low temperature impact strength specification;

- fracture appearance transition or nil ductility temperature specification for impact testing;

- reduce carbon content to less than 0.4 percent; - use of higher nickel alloy steels;

- lubricant problems,

3.4 System analysis

The system of connected rotating parts must be compatible, free from critical speeds, torsional or other types of vibration, within the specified operating speed range no matter how induced. The enclosed gear drive designer or manufacturer is not responsible for this analysis, unless agreed to in the purchase contract.

4 Unit rating

Historically, many terms have been used to denote conditions of operations - both calculated and

actual. These terms have resulted in confusion as to the actual capability of the enclosed drive. Examples of terms previously used to denote some form of enclosed drive capacity are listed below:

- service rating; - nameplate rating; - equivalent rating; - catalog rating; - mechanical rating; - brake rating; - unity rating; - transmitted horsepower; - calculated horsepower; - allowable horsepower; - application horsepower.

For purposes of this standard, where component capacities are being determined, the calculations are specifically related to the unit rating as defined below.

4.1 Unit rating definition

The unit rating is the overall mechanical power rating of all static and rotating elements within the enclosed drive. The m inimum rated component (weakest link, whether determined by gear teeth, shafts, bolting, housing, etc.) of the enclosed drive determines the unit rating.

4.2 Unit rating requirements

The unit rating implies that all items within the gear drive have been designed to meet or exceed the unit rating. Gear and pinion ratings are to be in accordance with the bending strength and pitting resistance ratings as outlined in this standard.

Shaft stresses, key stresses and fastener stresses are to be within the lim its set by this standard. Rolling element bearing or sleeve bearing designs are to be within lim its set by ANSIIAGMA 6001 -D97. Where user requirements or specifications dictate different design criteria, such as higher bearing life, this must be by contractual agreement.

Unit ratings may also include allowable overhung load values which are usually designated to act at a distance of one shaft diameter from the face of the housing or enclosure component. Stresses in related parts resulting from these overhung loads must also be within lim its set by this standard. Refer to clause 7 for further information.

5

ANSI/AGMA 6010-F97 AMERICAN NATIONAL STANDARD

4.3 Application of unit rating

The required unit rating of an enclosed drive is a function of the application and assessment of variable factors that affect the overall rating. These factors include environmental conditions, severity of service and life. Refer to clause 8 for further explanation.

The application of the enclosed drive requires that its capacity as defined by its unit rating; i.e., its m inimum rated component power, Pmcr be related to the actual service conditions.

..,

where

- alignment and deflection;

- bearing type and assembly;

- seals; fi is the application power of enclosed drive, - shaft driven accessories such as fans and

hp; pumps. P mc is the m inimum component power rating, hp;

K$ is the service factor.

For cases where the gear mesh has been deter- m ined to be the m inimum rated component, the lesser of Pa, or Pat, as defined in this standard, becomes Pmc in equation 1 (see 5.1 and 5.2).

4.4 Momentary overloads

When the enclosed drive is subjected to infrequent momentary overloads, stall conditions and low- cycle fatigue (less than 100 cycles), the conditions should be evaluated to assure that the yield strength of any component is not exceeded (see 8.3.1).

The heat dissipation characteristics and interaction of these factors are complex. The conditions of this standard prevail where rolling element bearings and hydrodynamic lip seals are used and a nominal gear efficiency value of 98 percent per mesh may be assumed. When other components (and their associated setups) are required, such as bearing preloads, face seals, high oil levels, special gear geometry, journal bearings and shifting mecha- nisms, a detailed analysis is necessary. It is beyond the scope of this standard to present a detailed analysis of efficiency.

With respect to the gear bending strength for momentary overloads, the maximum allowable stress is determined by the allowable yield properties rather than the bending fatigue strength of the material. This stress is designated as say; its determination is shown in ANSVAGMA 2001 -C95. Shaft, bearing and housing deflections have a significant effect on gear mesh alignment during momentary overloads. The enclosed drive must be evaluated to assure that the reactions to momentary overloads do not result in excessive m isalignment causing localized high stress concentrations or permanent deformation or both. In addition, the effects of external loads such as overhung, transverse and thrust loads must be evaluated.

An estimate of the efficiency of an enclosed drive may be made by using the thermal capacity calculated in clause 6, but such an estimate will be only an approximation of efficiency under test. The temperature sensitive portions of power loss may be different from those calculated according to clause 6, unless the operating sump temperature is nearly 200°F. Unless specifically agreed to between the user and manufacturer, items such as the prime mover, couplings, external driven loads, attaching devices and motor driven accessories, are not included in the enclosed drive efficiency estimate.

5 Rating criteria

4.5 Efficiency estimate

There are certain applications where the efficiency of an enclosed drive must be estimated and its thermal rating determined.

The pitting resistance power rating and the bending strength power rating for each mesh in the unit must be calculated and the lowest value obtained shall be used as the power rating of the gearset. It is permissible to use more conservative values.

The determination of efficiency is dependent on many complex characteristics and relationships. There are many factors which affect efficiency values. Some factors are:

- operating temperature; - ambient temperature;

- load and speed; - gear geometry; - lubricant and lubrication system; - housing characteristics;


5.1 Pitting resistance power rating, Pat

The pitting of gear teeth is considered to be a Hertzian contact fatigue phenomenon. Initial pitting and destructive pitting are illustrated and discussed in ANSVAGMA lOlO-E95, Appearance of Gear Teeth - Terminology of Wear and Failure.

conservative approach such as curves 1 or 2 is allowed. if assembled quality level is unknown, curve 1 should be used.

5.1.1.3 Stress cycle factor, &7

This factor adjusts the rating of individual gear elements based on the relative number of subjected stress cycles. This does not in anyway imply a fixed life. It adjusts each gear element rating based on the relative number of cycles. The number of cycles corresponding to 10 000 hours should be used to determine the ZN factor.

The purpose of the pitting resistance formula is to determine a load rating at which destructive pitting of the teeth does not occur during their design life. The ratings for pitting resistance are based on the formulas developed by Hertz for contact pressure between two curved surfaces, modified for the effect of load sharing between adjacent teeth.

5.1.1 Pitting resistance of spur and helical gears

The pitting resistance power rating shall be per the rating procedures and formulas of ANSVAGMA 2001 -C95. The following factors for enclosed drives shall be used:

KT = 1 .O, temperature factor;

Cf = 1 .O, surface condition factor;

In figure 17 of ANWAGMA 2001 -C95, for stress cycle factors above 1 x lo7 cycles, use the upper curve for pitting.

5.1.2 Pitting resistance of bevel gears

The pitting resistance power rating for bevel gears shall be per the rating procedures and formulas of ANSVAGMA 2003-A86. The following factors for enclosed drives shall be used:

CS = 1 .O, size factor;

&7 = 1 .O, reliability factor;

SH = 1 .O, safety factor;

&, is the dynamic factor (see 5.1 .l .l);

K, is the load distribution factor (see 5.1 .1.2);

K7 = 1 .O, overload factor;

ZN is the stress cycle factor (see 5.1.1.3).

5.1 .l .l Dynamic factor, &

K,, is to be determined by clause 8 of ANSVAGMA 2001 -C95, with the exception that values of I$ shall not be based on a transmission accuracy number greater than QV = 11.

Transmission accuracy number (Q,,) shall be based upon the quality level of the gearing produced. In the absence of a known specific quality level, the curve

Cf = 1 .O, surface condition factor;

G = 1 .O, external dynamic factor;

cb = 0.735, stress adjustment factor;

CL = 1 .O, life adjustment factor;

C,, is the dynamic factor (see 5.1.2.2);

Gl is the load distribution factor (see 5.1.2.3);

CT = 1 .O, temperature factor.

5.1.2.1 Life factor (stress cycle factor), CL

This factor adjusts the rating of individual gear elements based on the relative number of load cycles. The number of cycles corresponding to 10 000 hours should be used to determine CL. See figure 1 in ANSVAGMA 2003-A86. The upper curve should be used.

for Q , = 6 shall be used. 5.1.2.2 Dynamic factor, C,,, &

Figure 1 of ANSVAGMA 2001 -C95 lim its values for C,, and & are to be determined by clause 8.5 in the dynamic factor as a function of pitch line velocity. ANSVAGMA 2003-A86, with the exception that These curves cannot be extrapolated to obtain values of C,, and II; shall not be based on a values beyond the lim its given. transmission accuracy number greater than Q, = 11.

5.1 .1.2 Load distribution factor, & S, is set to the smaller of the gear or pinion allowable stress numbers, S,,.

The empirical method of ANSVAGMA 2001-C95 shall be used in determining the load distribution factor. For typical enclosed drives, the mesh alignment factor, C,, , shall be obtained from curve 3 of figure 7 in ANSVAGMA 2001-C95. A more

5.1.2.3 Load distribution factor, C,, If&

C, and I& are to be determined by 10.1 of ANSVAGMA 2003486 by setting design pinion torque, To, equal to operating pinion torque, Tp.

7


5.2 Bending strength power rating, Pat

Bending strength of gear teeth is a measure of the resistance to fatigue cracking at the tooth root fillet. Typical cracks and fractures are illustrated in ANSI/ AGMA 101 O-E95

The intent of the AGMA strength rating formula is to determine the load which can be transmitted for the design life of the teeth without causing root fillet cracking or failure.

Occasionally manufacturing tool marks, wear, surface fatigue or plastic flow may lim it bending strength due to stress concentration around large, sharp cornered pits or wear steps on the tooth surface.

5.2.1 Bending strength of spur and helical gears

The bending strength rating for gearing within the scope of this standard shall be determined by the rating methods and procedures of ANSI/AGMA 2001 -C95. The following factors for enclosed drives shall be used:

KT = 1 .O, temperature factor; Iij, = 1 .O, reliability factor;

SF = 1 .O, safety factor; Iyy is the dynamic factor (see 5.1 .l .l);

K, is the load distribution factor (see 5.1 .1.2);

& = 1 .O, overload factor;

YN is the stress cycle factor (see 5.2.1 .l).

5.2.1 .l Stress cycle factor, YN

This factor adjusts the rating of individual gear elements based on the relative number of subjected stress cycles. This does not in anyway imply a fixed life. It adjusts each gear element rating based on the relative number of cycles. The number of cycles corresponding to 10 000 hours should be used to determine the YN factor.

& is the dynamic factor (see 5.1.2.2);

& = 1 .O, temperature factor; & is the stress cycle factor (see 5.2.2.1); K, is the load distribution factor (see 5.1.2.3).

5.2.2.1 Life factor (stress cycle factor), & This factor adjusts the rating of individual gear elements based on the relative number of load cycles. The number of cycles corresponding to 10 000 hours should be used to determine &. See figure 2 in ANWAGMA 2003-A88. The lower curve shall be used. 5.2.2.2 Momentary starting loads Since the bending strength rating practice for bevel gears of ANSI/AGMA 2003-A86 does not account for momentary peak loads as encountered during starting, a load spectrum analysis, such as with M iner’s Rule, is required to account for the permissible starting and operating peak load cycles.

5.3 Allowable stress numbers for pitting resistance and bending strength As defined in the gear tooth rating standards, the allowable stress numbers for gear materials vary with composition, cleanliness, quality, heat treat- ment and processing practices. Unless justified by testing, do not use the allowable stress numbers for Grade 3 material. Use the allowable stress numbers for Grades 1 or 2 in accordance with ANWAGMA 2001 -C95, clause 16, or ANSI/AGMA 2003-A86, clause 20.

6 Thermal power rating

The following thermal model has been established using empirical factors. It is based on the experience of several gear manufacturers. The model has been validated by extensive testing of concentric shaft, base mounted reducers with shafts mounted in a horizontal orientation. Limited testing of some parallel shaft gear units has also been performed to spot check the adequacy (validity) of the model. Values of some variables such as arrangement constant, heat transfer coefficient and coefficient of friction may not adequately address other enclosed drive configurations and operating conditions. These configurations or conditions may necessitate modifications of these variables. Changing any variable requires care and testing to insure that the principles of the heat balance formulation are not violated.

In figure 18 of ANSI/AGMA 2001 -C95, for stress cycle factors above 3 x lo6 cycles, use the upper curve for bending.

5.2.2 Bending strength of bevel gears

The bending strength rating of bevel gears shall be determined by the rating methods and procedures of ANWAGMA 2003486. The following factors for enclosed drives shall be used:

& = 1 .O, external dynamic factor;

49 = (2.8 - Pd”-25)/1.2, size factor;

8

AMERICAN NATIONAL STANDARD ANSI/AGMA 6010~F97

Maintaining an acceptable temperature in the oil sump of a gear drive is critical to its life. Therefore, the selection of a gear drive must consider not only the mechanical rating but also the thermal rating.

Thermal rating is defined as the maximum power that can be continuously transmitted through a gear drive without exceedina a soecified oil sump temperature. The thermal rating must equal or exceed the transmitted power. Service factors are not used when determining thermal requirements. The thermal rating depends upon the specifics of the drive, operating conditions, the maximum allowable sump temperature, as well as the type of cooling employed.

6.2 Service conditions

6.2.1 Intermittent service

For intermittent service, the input power may exceed the manufacturer’s thermal power rating provided the oil sump temperature does not exceed 200” F.

6.2.2 Adverse conditions

The ability of a gear drive to operate within its thermal power rating may be reduced when adverse conditions exist. Some examples of adverse environmental conditions are:

_ an enclosed space;

- a buildup of material that may cover the gear drive and reduce heat dissipation;

6.1 Rating criteria - a high ambient temperature, such as boiler,

The primary thermal rating criterion is the maximum allowable oil sump temperature. Unacceptably high oil sump temperatures influence gear drive operation by increasing the oxidation rate of the oil and decreasing its viscosity. Reduced viscosity translates into reduced oil film thickness on the gear teeth and bearing contacting surfaces which may result in reducing the life of these elements. To achieve the required life and performance of a gear drive, the operating oil sump temperatures must be evaluated and lim ited.

Thermal ratings of gear drives rated by this standard are lim ited to a maximum allowable oil sump temperature of 200” F. However, based on the gear manufacturer’s experience or application requirements, selection can be made for oil sump temperatures above or below 200°F (see 6.6).

Additional criteria that must be applied in establishing the thermal rating for a specific gear drive with a given type of cooling are related to the operating conditions of the drive. The basic thermal rating, PT, is established by test (Method A) or by calculation (Method B) under the following conditions:

machinery or turbine rooms, or in conjunction with hot processing equipment; - high altitudes; - the presence of solar energy or radiant heat.

6.2.3 Favorable conditions

The thermal power rating may be enhanced when operating conditions include increased air movement or a low ambient temperature.

6.2.4 Auxiliary cooling

Auxiliary cooling should be used when the thermal rating is insufficient for operating conditions. The oil may be cooled by a number of means, some of which are:

- Fan cooling. The fan shall maintain the fan cooled thermal power rating; - Heat exchanger. The heat exchanger used shall be capable of absorbing generated heat that cannot be dissipated by the gear drive by convection and radiation.

6.3 Methods for determining the thermal rating

Thermal rating may be determined by one of two methods: Method A - test, or Method B - calculation.

6.3.1 Method A: Test

Test of full scale gear drives at operating conditions is the most accurate method for establishing the thermal rating of the gear drive. See 6.4.

6.3.2 Method B: Heat balance calculation

- oil sump temperature at 200°F;

- ambient air temperature of 75°F;

- ambient air velocity of 5275 fpm in a large indoor space;

- air density at sea level; and

- continuous operation. The thermal rating of a gear drive can be calculated using the heat balance equation which equates heat generated with heat dissipated. The method for

Modifying factors for deviation from these criteria are given in 6.6.

9


calculating the thermal rating is given in 6.5. The The basis of the thermal rating is when the losses, method for calculating heat generation is discussed Pv, at PA, are equal to the heat dissipation, PQ, of the in 6.51 and for heat dissipation in 6.5.2. gear drive.

6.4 Method A - Test

A test of a specific gear drive at its design operating conditions is the most reliable means to establish the thermal rating. Thermal testing involves measuring the steady state bulk oil sump temperature of the gear drive operating at its rated speed at no load and at least one or two increments of load. Preferably one test should be at 200°F sump temperature.

While no load testing cannot yield a thermal rating, it may be used to approximate the heat transfer coefficient for comparison purposes, if the power required to operate the drive at no load is measured.

Some guidelines for acceptable thermal testing are as follows:

- The ambient air temperature and velocity must be stabilized and measured for the duration of the test;

- The time required for the gear drive to reach a steady state sump temperature depends upon the drive size and the type of cooling;

- Steady state conditions can be approximated when the change in oil sump temperature is 2°F or less per hour.

The oil temperature in the sump at various locations can vary as much as 27°F. The location of the temperature measurement should represent the bulk oil temperature. Outer surface temperatures can vary substantially from the sump temperature. The opposite direction of rotation can create a different sump temperature.

PQ =PV . ..(2) When this is satisfied under the conditions of 6.1, input power, PA is equal to the thermal power rating, PT.

The heat generation in a gear drive comes from both load dependent, PL, and non-load dependent losses, PN.

PV =PL +PN PL is a function of the input power, PA.

. ..(3)

PL = f(PAA) . ..(4)

Using equation 2 and rearranging terms, we can write the basic heat balance equation as follows:

PQ -PN-f(a) =o . ..(5) To determine the basic thermal rating, PT, vary PA until equation 5 is satisfied. This can be done by recalculating the load dependent losses, PL, at different input powers, PA. if PQ s PN, the gear drive does not have adequate thermal capacity. The design must be changed to increase PQ or auxiliary cooling methods must be used.

When equation 5 is satisfied, the overall unit efficiency, q, is calculated as follows:

q = loo- ‘LipN PA

x 100 . . .

The thermal rating of the gear drive can be related to efficiency as follows:

6.51 Heat generation The heat generated in a gear drive comes from both load dependent, PL, and non-load dependent

During thermal testing the housing outer surface temperature can be surveyed if detailed analysis of the heat transfer coefficient and effective housing surface area is desired. Also, with fan cooling, the air velocity distribution over the housing surface can be measured.

6.5 Method B - Calculations for determining the thermal power rating, PT

losses, PN. The load dependent losses are comprised of the bearing losses, PB, and the gear mesh losses, PM:

pL = cpB + CPM . ..(8)

The non-load dependent losses consist of the oil seal losses, Ps, the internal windage and oil churning losses, Pw and qyg, and the oil pump power, Pp, consumed.

The calculation of thermal rating is an iterative process due to the load dependency of the coefficient of friction for the gear mesh and the bearing power loss.

PN = -& + -&J + CPw + &I . ..(8)

These losses must be summed for each occurrence in the gear drive.

10


6.5.1 .l Bearing power loss, PB

Rolling contact bearing power loss, PB, may be estimated by using equations 10 and 11. Values for the bearing coefficient of friction, j,, may be approximated using the values from table 2 [l]. When more exact values are known, they should be used. For more detailed information see [l], [2], [3] and 141.

churning effects have often been combined with the assumed friction values. Ideally, the coefficient of friction depends on the lubricant properties, surface conditions and sliding velocity. It also changes with contact load factor, K. 6.5.1.2.1 Mesh power loss, PM, spur and helical gears For spur and helical gears, the following equation can be used to estimate the gear tooth mesh losses 111, [51, PI and Fl:

Tb nb -- 'B - 63025

. . .

where

Tb is the rolling bearing friction torque, lb in;

T b

= fbw(do + di) 4

. ..(n)

nb is bearing shaft speed, rpm;

I% is bearing coefficient of friction (table 2);

w is bearing load, lb;

4 is bearing outside diameter, in;

4 is bearing bore, in.

Table 2 - Bearing coefficient of friction,fb Coefficient

Type of bearing of friction’),& Radial ball bearing

(single-row deep groove) 0.0015 Self-aligning ball bearing 0.0010 Angular-contact ball bearing 0.0013 Thrust ball bearing 0.0013 Cylindrical roller bearing 0.0011 Spherical roller bearing*) 0.001 a Tapered roller bearing*) 0.001 a NOTE: 1) Variation in fb depends on speed and load. 2) j, is greater on tapered and spherical roller bearings due to rubbing on the roller ends.

6.5.1.2 Mesh power loss, PM

Mesh losses are a function of the mechanics of tooth action and the coefficient of friction. Tooth action involves some sliding with the meshing teeth separated by an oil film .

The mesh efficiency is expressed as a function of the specific sliding velocities and the coefficient of friction.

The coefficient of friction is difficult to assess. Reliable published data is rather lim ited, especially at high pitch line velocities. In the past, windage and

p M

= fnlTpj?~s”%v 63025M

. ..(12)

where

fm is the mesh coefficient of friction at mesh oil temperature;

If the pitch line velocity, V, is 400 < v < 5000 fpm and the contact load factor, K, is 100 c K < 2000 lb/in* and IS0 VG is between 46 and 460, then fm can be estimated by equation 13. Outside these lim its the mesh coefficient should be determined experimentally.

fi.35 fm = -

c,vo= . . . '(13)

where

Tp is the torque on the pinion, lb in;

np is the rpm of the pinion; wy is the operating helix angle at operating

pitch diameter, deg; M is the mesh mechanical advantage; Cl is lubricant factor (see table 3); V is pitch line velocity, fpm.

Table 3 - Lubricant factor, Cl, at 200°F sump

AGMA 1 2 4

I 5 6 7

NOTE:

temperature’) IS0 VG

46 68 150 220 320 460

Cl 101.8 95.3 84.6 80.1 76.3 72.6

1) These values can be approximated bq Cl= 172.85 x (IS0 VG)-".142.

K is given by the equation:

K _ TP @ P + a

2~ (rd2 NC . ..(14)

f 1

11

ANSIJAGMA 6010-F97

where

Np is the number of pinion teeth;

NG is the number of gear teeth;

F is the face width in contact with mating element, in;

rw is the pinion operating pitch radius, in.

The equation for the mesh mechanical advantage is:

where

M= 2~s~w(H,+~J . . fe+* #.(15)

where

Cp, is the operating transverse pressure angle, deg;

w, is the sliding ratio at start of approach;

& is the sliding ratio at end of recess.

The values for Hs and Ht are:

Hs = (??lG + l)[(j$-&w~-Sin@w]

. ..(16)

Ht = (w)[ (-$cos2+w~-sin~w] where

. ..(l?)

mu is the gear ratio, NG/NP;

I& is the gear outside radius, in;

R,,, is the gear operating pitch radius, in;

r. is the pinion outside radius, in.

6.5.1.2.2 Mesh power loss, PM, bevel gears

A method for determining bevel gear mesh power loss is shown in annex F.

6.5.1.3 Oil seal power loss, Ps

Contact lip oil seal losses are a function of shaft speed, shaft size, oil sump temperature, oil viscosity, depth of submersion of the oil seal in the oil and oil seal design. Oil seal power losses can be estimated from equation 18. Figure 1 can be used to estimate oil seal frictional torque as a function of shaft diameter for oil seals typically used in gear drives, see [8].

TSn ‘s = 63 025 418) . . .

Ts is the oil seal torque, lb in (figure 1); n is the shaft speed, rpm.

I I I x I I I l/l l/f- ”

iii P* .9 6 1 5 CO

0 1 Sh% diaieter,4DS, 5 6 7 in

Figure 1 - Seal friction torque

6.5.1.4 Gear windage and churning power loss, pw For gear drives covered by this standard, windage and churning losses are generally combined into a single loss. This loss, Pw, for each gear and pinion can be estimated from equation 19 and 20 for spur or helical gears, and equations 21 and 22 for bevel gears. The empirical arrangement constant, A, varies with the arrangement of the gears in the gear drive, the degree of contact with the oil and the oil viscosity. The same arrangement constant can be used for gears, pinions and bearings. For gear drives covered by this standard, the arrangement constant is given by equation 23.

‘WG = d2, n2 F, cd qw

126 000 Pn A ,..(19)

P u?P= d2, n2 Ft cm3 tjw . . .

126 000 P, A

pwG = Pm)* n* Ft cos3 v 126 000 P,,,,, A

... (20)

421)

P (2r,J2 n* Ftcos3 IJJ

WP’ 126 000 Pm,, A . ..(22)

where

PWG is the windage loss for gear; Pv is the windage loss for pinion;

44 is the operating pitch diameter of gear or pinion, in;

n is the shaft speed (gear or pinion), rpm;

Ft is the total face width of gear or pinion, in;

‘1c1 is the mean spiral angle, degrees;

h is the mean reference radius, in;

12


ptl is the normal diametral pitch, in-‘; required power is a function of the oil flow and oil

P mn is the mean normal diametral pitch, in-‘; pressure at a given pump speed.

A is the arrangement constant. PP = PPS + PPnl . ..(27) For an oil pump driven by one of the reducer shafts,

A _ 22440 -- CP

the oil pump loss, Pps, can be estimated by equation -(23) 28

where

CP is the absolute oil viscosity at sump PPS = g$ . . . (28)

temperature, cP.

6.5.1.5 Bearing windage and churning power loss, *

For gear drives covered by this standard, windage and churning losses are generally combined into a single loss. For bearings other than tapered roller bearings, the windage and churning losses are included in PB. This is a simplified approach combining no load and load dependent losses.

where

Q is oil flow, gal/min;

P is operating oil pressure, lb/in*;

ep is oil pump efficiency, usually around 85%.

For an oil pump driven by an electric motor, the oil pump losses, Pp,, can be estimated using equation 29, which considers the electric power consumed and the efficiencies of both the electric motor and the oil pump.

For tapered roller bearings, this loss, Pm, can be estimated for each bearing from equation 24.

For tapered roller bearings only: PPI?l = EP . ..(29)

where

P,= d2, n2 B coo83 aB D, . .

126 000 rr (0.78) A ..(24) Ep is electric power consumed, hp;

pm is electric motor efficiency, %. where

n is shaft speed (each bearing), rpm;

B

42

is length thru bore of bearing, in;

is mean bearing diameter (l/2 (bearing cup

DR is the mean roller diameter, in;

outer diameter + bearing cone bore

Q3

diameter)), of the tapered roller bearing, in;

is cup angle of a tapered roller bearing.

. ..(25)

The value of e is determined from the bearing manufacturer for the specific bearing number, or when e is not provided,

The heat dissipated from a gear drive is influenced

Ppm should be included in the thermal calculations,

by the surface area of the gear drive, the air velocity across the surface, the temperature differential, M, between the oil sump and the ambient air, the heat transfer rate from the oil to the gear case and the

but should not be included in the enclosed gear unit

heat transfer rate from the gear case and the ambient air. The heat dissipation is given by

efficiency. See 4.5.

equation 30.

6.5.2 Heat dissipation, PQ

P,=A,ku . ..(30) where

QB = tm--l 0.389

( 1 K5 . . .

where

K5 is the ratio of basic dynamic radial load rating to basic dynamic thrust load rating.

The value of KS is available from the bearing manufacturer for the specific bearing number.

6.5.1.6 Oil pump power loss, Pp

A, is the gear case surface area, ft*;

k is the heat transfer coefficient, hp/ft* “F (see table 4 or 5);

AT is the temperature differential, “F. NOTE: & is the gear case surface area exposed to ambient air, not including fins, bolts, bosses or mounting surfaces. CAUTION: The lubricant must be selected to accommodate the extreme conditions of the temperature differential. See clause 9.

The required power and capacity of most lubrication The heat transfer coefficient, k~, is defined as the oil pumps vary directly with the speed. Thus, the average value over the entire gear drive outer

13

ANSIIAGMA 6010-F97 AMERICAN NATIONAL STANDARD

surface. The heat transfer coefficient will vary depending upon the material of the gear case, the cleanliness of the external s&ace, the extent of wetting of the internal surfaces by the hot oil, the configuration of the gear drive and the air velocity across the external surface. For gear drives covered by this standard, typical values for k can be found in table 4. See [9].

6.6 Corrections for non-standard operating conditions

When the actual operating conditions for a specific application are different from the standard conditions defined in 6.1, and the thermal rating is calculated for the conditions of 6.1, the thermal rating may be modified for the application as follows:

PThm =pT&ef&B~ BTBD . ..(31)

Bref and BA may be applied to natural or shaft fan cooling. BV may be applied only to natural cooling.

Table 4 - Heat transfer coefficient, k, for gear drives without auxiliary cooling

Condition Small confined space

Air velocity,

fpm ~275

Heat transfer coefficient, kl)

hp/ft2 “F 0.0007 - 0.0010

Large indoor space

5275

>275

>725

0.0011 - 0.0014

0.0012 - 0.0015

0.0014 - 0.0017

Large indoor space Outdoors NOTE: ‘1 The choice of kvalues within each range is affected by the items listed in 6.5.2. Use of the high values in each range should be justified by test.

The heat transfer coefficient for a shaft fan cooled gear drive is a function of fan design, shroud design and fan speed. It will vary substantially depending upon the effectiveness of the fan and the proportion of the exterior surface cooled by the resulting airflow. The air velocity is defined to be the average air velocity over 60% of the surface area, 4, of the gear drive. The effect of using multiple fans on a gear drive could increase the average air velocity, thereby resulting in a higher heat transfer coefficient. Table 5 provides values for k for fan cooled gear drives.

Table 5 - Heat transfer coefficient, k, for gear drives with fan cooling

Air velocity Heat transfer coefficient, k fm hplft2 “F 500 0.0010 1000 0.0017 2000 0.0029 3000 0.0040

The gear drive manufacturer should be consulted when the conditions exceed the limits given in tables 6 thru 10 or when correction factors are required for any type of cooling other than natural or shaft fan.

When the ambient air temperature is below 75”F, Bref allows an increase in the thermal rating. Conversely, with an ambient air temperature above 75”F, the thermal rating is reduced. See table 6.

Table 6 - Ambient temperature modifier, Bref

Ambient temperature, “F hef

ii2 1.15 1.07

75 1.00 85 0.93

100 0.63 110 0.75 120 0.67

When the surrounding air has a steady velocity in excess of 275 fpm due to natural or operational wind fields, the increased convection heat transfer allows the thermal rating to be increased by applying Bv. Conversely, with an ambient air velocity of 1100 fpm, the thermal rating is reduced. See table 7.

Table 7 - Ambient air velocity modifier, BV

I~

At high altitudes the decrease in air density results in the derating factor, BA. See table 8.

14


Table 8 - Altitude modifier, BA

Altitude, ft BA 0 - Sea level 1 .oo

2500 0.95 5000 0.90 7500 0.85

10 000 0.81 12 500 0.76 15 000 0.72 17500 0.68

The standard maximum allowable oil sump temperature is 200°F. A lower sump temperature requires a reduction in the thermal rating using BT. See table 9. A maximum allowable sump temperature in excess of 200°F will increase the thermal rating and can provide acceptable gear drive performance in some applications. However, it must be recognized that operating above 200°F may reduce lubricant and contact seal life and increase the surface deterioration on the gears and bearings with a subsequent increase in the frequency of maintenance. The gear manufacturer should be consulted when a maximum allowable oil sump temperature in excess of 200°F is being considered.

Table 9 - Maximum allowable oil sump temperature modifier, BT

When a gear drive sees less than continuous operation with periods of zero speed, the resulting “cool-off” time allows the thermal rating to be increased by&. See table 10.

Tabl;! 10 - Operation time modifier, BD

Operation time per each hour

100% (continuous) 80% 60% 40% 20%

BD

1.00 1.05 1.15 1.35 1.80

7 Component design

The components of a gear drive must be designed with consideration for all loads likely to be encountered during operation. These include the torque loads imposed. on the components through the gearing, and the external loads, such as overhung loads, external thrust loads and dynamic loads. Components must also be designed to withstand any assembly forces which m ight exceed the operating loads. During the design process, the operating loads must be considered to occur in the worst possible direction and loading combinations.

All components shall allow for peak loads of 200 percent of the unit rating, considering both internal and external loads, in accordance with 3.1. User requirements or specifications dictating different design criteria must be by contractual agreement.

7.1 Housing

Refer to clause 7 of ANSVAGMA 6001-D97 for design guidance.

7.2 Bearings

Shafts may be mounted in sleeve or rolling element bearings, of a size, type and capacity to carry the radial and thrust loads that would occur under maximum operating conditions. For additional information, consult clause 6 of ANSVAGMA 6001 -D97.

7.2.1 Sleeve bearings

Sleeve bearings shall be designed for maximum bearing pressures of 750 lb/in* on projected area. Journal velocities shall not exceed the values given below:

- 1500 fpm with lubricant supplied not under pressure; - 7200 fpm with lubricant supplied under gravity with the oil inlet fully flooded.

7.2.2 Roller and ball bearings

Roller and ball bearings shall be selected to provide a m inimum Lfc bearing life of 5000 hours based on unit rating as calculated by the methods of the bearing manufacturers, with considerations given to lubrication, temperature, load zone, alignment and bearing material.

7.3 Shafting

Shafting should be designed in accordance with clause 4 of ANSVAGMA 6001 -D97.

15

sSO = s& x (100 - T, x I&) . ..(33)

100 where

$0

%O

s,

is the adjusted compressive stress, lb/in*;

is the adjusted shear stress, lb/in*;

is the calculated key compressive stress, lb/in*;

is the calculated key shear stress, lb/in*;

is the allowable percentage of torque capacity obtained due to the interference fit (by calculation or test);

is the shrink adjustment factor to compensate for less than 100 percent engagement (determined by test or experience).

ANSI/AGMA WlO-F97

7.4 Keys

Refer to clause 5 of ANSI/AGMA 6001-D97. See annex 6.

7.4.1 External keyways

Keyways in external shaft extensions on the gear drive should conform to ANSI 817 “commercial class” or IS0 R773-1969 (E) “free fit”.

7.4.2 Allowance for interference fit

The allowable stresses provided in annex A of ANSI/AGMA 6001 -D97 are based on the assumption that an interference fit is not used and that the key carries the entire torque load. When an interference fit is used in conjunction with a key, the actual compressive or shear stress may be reduced by the effect of the interference fit at maximum operating temperature as follows:

SC0 = SC x ( 100 - Tf x Ksa 1 . ..(32) 100

The product of Tf& 5 100.

7.5 Threaded fasteners

Refer to clause 8 of ANSVAGMA 6001 -D97.

7.6 Backstops

Backstops are designed to prevent reverse rotation of driven equipment that is intended for uni- directional rotation only. They allow free, unimpeded rotation in one direction, while preventing rotation in the opposite direction. Specified torque limit and rotational speed will vary depending upon the manufacturer.

7.6.1 Types

Backstops are a variety of clutch. This discussion is limited to cam or “sprag” type clutches. However, other types may be used. Cam clutches are generally used for three distinct operational modes: overrunning, indexing and backstopping. This discussion is further limited to the backstopping mode of operation.

7.6.2 Selection and application

Backstopping load is permitted to pass through all components between the load and the backstop. This may render the backstop function ineffective in case of component failure between the backstop and the driven load.

The maximum allowable overrunning backstop speed must be greater than the maximum shaft speed attainable in all operating conditions.

The backstop should be selected based upon the number of backstopping cycles and the applied torque.

7.6.3 Installation

A backstop is installed with the outer race of the backstop anchored to a stationary member while the inner members can overrun freely in one direction of rotation. A backstop with a separate inner race must have the race secured to the rotating shaft.

7.6.3.1 Installation of built in types

A backstop can be built into the gear drive and the inner race eliminated by having the cams engaging directly on the supporting shaft. In this case, the inner race surface must be capable of sustaining a Hertzian contact stress as required by the backstop design.

The following items concerning the shaft must also be addressed:

- surface hardness; - case depth; - taper of the supporting cam contact area; - concentricity with backstop outer race.

The shaft that the backstop is mounted on must be supported by bearings. The backstop is not intended to withstand reaction loads imposed by gears or other sources.

7.6.3.2 Source of installation

Some manufacturers choose to ship gear drives with backstops internal to the gear drive while others

16

AMERICAN NATIONAL STANDARD ANSI/AGMA 6010-F97

make them external. In either case, special care must be taken to identify the proper direction of rotation. Manufacturers usually identify the backstop housing with a marking to indicate the appropriate shaft rotation to prevent damage to the backstop. The installation should take the added precaution of manually rotating the input shaft with the motor installed to insure the proper operation of the clutch.

7.6.4 Lubrication

The backstop, much like other elements of the gear drive, must have proper lubrication that is free of contaminants, and the lubricant should be drained, flushed, and changed on a regular maintenance schedule as recommended by the gear drive manufacturer.

The backstop may be lubricated by grease or oil that is appropriately selected for the application considering the environment in which it will operate.

WARNING: Do not use extreme pressure lubricants or lubricants with formulations including sulphur, chlorine, lead and phosphorous derivatives, as well as graphite and molybdenum disulfide in gear drives equipped with an internal backstop unless approved by the gear drive manufacturer or the backstop manufacturer. WARNING: Some synthetic gear lubricants adversely affect the operation of internal backstops. Special authorization is required from the gear drive manufacturer before using a synthetic lubricant in a gear drive equipped with an internal backstop.

7.7 Balancing

The purpose of balancing is to minimize or eliminate vibration in a rotating element due to unbalance. The importance of proper balancing increases directly in proportion to the pitch line velocity of the rotating part. Excessive unbalance can result in premature bearing, gear or other component failure.

It is the responsibility of the manufacturer of the drive components to determine the need for balancing and assure that it is done without affecting the structural integrity of the rotating mass.

7.8 Shrink discs

Shrink discs are one option to connect a hollow shaft gear drive to the drive shaft. The shrink disc is an external locking device installed over a hollow shaft projection. By tightening the locking screws, the locking collars exert radial forces on the tapered inner ring and the hub. After bridging the fit clearances, radial clamping pressure is generated between the drive shaft and the hollow shaft establishing a solid, frictional connection.

During the selection process, the following items should be considered:

- hollow shaft and drive shaft yield point of material;

- coefficient of friction between the hollow shaft and the drive shaft;

- tolerance and fits of the mating surfaces;

- effect of diameter change when the shrink disc is applied;

- surface finishes of the hollow shaft and drive shaft;

- axial forces applied to the assembly;

- starting and peak loads transmitted through the drive system.

The shrink disc should be selected according to the manufacturer’s recommendations.

7.9 Other components

See clause 9 of ANSVAGMA 6001-D97 for brief discussions on the following components:

- shims, 9.1;

- gaskets, 9.2;

- oil seals, 9.3 (see note below);

- breathers, 9.4 (see note below);

- expansion chambers, 9.5 (see note below);

- oil level indicators, 9.6;

- bearing retainers, 9.7;

- grease retainers, 9.8;

- dowels and pins, 9.9;

- spacers, 9.10;

- seal retainers, 9.11;

- locking devices for fasteners, 9.12;

- tolerance and fits of mating surfaces.

NOTE: It is recognized that gear drives applied in certain industries and under certain atmospheric conditions should be equipped with special seals and breathers designed for those conditions. Examples are units installed in the dusty or corrosive atmospheres of chemical plants, cement mills and taconite processing plants. It is also recommended that units which are to be exposed to severe moisture and vapor laden atmospheres be equipped with moisture barrier seats and breathers. Some applications in wet locations subject to direct or indirect wash down may preclude the use of breathers, such as in the paper and food industries. In these cases, expansion chambers may be used.

17


8 Service factors

8.1 Selection of service factors

Before an enclosed gear drive can be selected for an application, an “equivalent power rating” must be determined. This is done by multiplying the specified transmitted power by the service factor. Since service factors represent the normal relationship between gear unit design power rating and the maximum potential transmitted power, it is suggested that the service factor be applied to the nameplate rating of the driven machine or prime mover, as applicable.

Manufacturer and user must agree upon which power, prime mover rating or driven machine requirements, should dictate the selection of the gear drive. It is necessary that the gear drive selected have a rated load capacity equal to or in excess of this “equivalent power rating”.

Service factor has been used to include the combined effects of ZN, YN, & and Ii;, in an empirically determined single factor. The mathemat- ical contribution of each of these factors has not been established. See ANSVAGMA 2001 -C95, clause 10. In the absence of more specific load data, a service factor, &F, shown in table A.2 of annex A, may be used.

8.2 Listing of service factors

The table of “Service Factors”, shown in annex A, has been developed from the experience of manufacturers and users of gear drives for use in common applications and has been found to be generally satisfactory for the listed industries when gears are rated using AGMA standards. It is suggested that service factors for special applications be agreed upon by the user and the gear manufacturer when variations of the tabulated value may be necessary.

This standard is based on the premise that the user is defining a catalog rating.

8.3 Determining service factors

Service factors may be selected from annex A or may be determined by an analytical method. listed below are some of the more important factors to be considered.

8.3.1 Operational characteristics

Some of the operational characteristics that could affect an increase or decrease in service factors are:

- Type of prime mover. Differenttypes of prime movers are electric motors, hydraulic motors, steam or gas turbines, and internal combustion engines having single or multiple cylinders.

- Starting conditions. Starting conditions where peak loads exceed 200 percent of rated load and frequency or duration as defined in 3.1. Rated load is defined as the unit rating with a servicefac- tor of 1.0.

When a soft start coupling is used between the prime mover and the gear drive, the selection of service factors can be based on the gear drive manufacturer’s analysis for the application.

- Overloads. Loads which are in excess of the rated load are considered overloads. Overloads can be of momentary duration, periodic, quasi- steady state, or vibratory in nature. The magnitude and the number of stress cycles require special analysis to prevent low cycle fatigue or yield stress failure.

Applications such as high torque motors, extreme repetitive shock, or where high energy loads must be absorbed, as when stalling, require special consideration.

- Overspeeds. Overspeeds contributing to external transmitted loads and dynamic loads require special analysis.

- Brake equipped applications. When a gear drive is equipped with a “working” brake that is used to decelerate the motion of the system, select the drive based on the brake rating or the transmitted power, whichever is greater. If the brake is used for holding only, and is applied after the motion of the system has come to rest, the brake rating should be less than 200 percent of the base unit rating. If the brake rating is greater than 200 percent of the unit rating, or the brake is located on the output shaft of the gear drive, special analysis is required.

- Reliability and life requirement. Applications requiring a high degree of dependability or unusually long life should be given careful consideration by the user and the gear manufacturer before assigning a service factor.

8.3.2 System conditions

An essential phase in the design of a system of rotating machinery is the analysis of the dynamic (vibratory) response of a system to excitation forces.

18


8.3.2.1 Vibration analysis

Any vibration analysis must consider the complete system including prime mover, gear drive, driven equipment, couplings and foundations. The dynamic loads imposed upon agear drive are the result of the dynamic behavior of the total system and not of the gear drive alone.

- Lubrication. Any lubricant not in accordance with manufacturer’s recommendations;

- Misalignment and distortions;

- Reversing applications;

- High risk applications involving human safety.

8.3.2.2 Dynamic response 9 Lubrication and lubricants

The dynamic response of a system results in additional loads imposed on the system and relative motion between adjacent elements in the system.

9.1 Lubrication

These lubrication recommendations apply only to enclosed gear drives which are designed and rated in accordance with current AGMA standards. Additional recommendations are contained within ANSVAGMA 9005D94. These recommendations are not intended to replace any specific lubrication recommendations made by the gear drive manufacturer.

The vibratory loads are superimposed upon the mean running load in the system and, depending upon the dynamic behavior of the system, could lead to failure of the system components.

8.3.2.3 System induced failure

In a gear drive, system induced failures could occur The lubricant must be selected to provide adequate as tooth breakage or severe surface deterioration of oil film thickness at all operating conditions. This the gear elements, shaft breakage, bearing failure or may require seasonal change of lubricant, oil failure of other component parts. heaters for cold starting conditions, or oil cooler for

8.3.2.4 Special system considerations high ambient temperatures. Oil film thickness is criiical to limit wear of gears and bearings.

It should be pointed out that synchronous motors, certain types of high torque induction motors and generator drives require special care in system design.

Synchronous motors have high transient torques during starting and when they momentarily trip-out and restart.

Induction motors of special high slip design can produce extremely high starting torques. Also, when the motor trips out for a very short time and then the trip re-closes, high torque loads are produced.

9.1 .l Ambient temperature

The ambient temperature range is -40” to 130°F and is defined as the air temperature in the immediate vicinity of the gear drive. Gear drives exposed to the direct rays of the sun or other radiant heat sources will run hotter and must therefore be given special consideration.

9.1.2 Other considerations

Gear drives operating outside of these temperature ranges, or those operating in extremely humid, chemical or dust laden atmospheres should be referred to the gear drive manufacturer. Generators have extremely high loads when they

are out of phase with the main system. Also, across-the-line shorts can produce torque loads up to twenty times the normal running torque.

9.1.3 Oil sump temperatures

The maximum oil sump temperature for mineral based oils is limited to 200°F. This sump temperature is considered maximum because many lubricants are unstable above the stated maximum

All special torque conditions should be considered when determining a service factor.

8.3.3 Special considerations temperature.

CAUTION: Sumb temoeratures in excess of 200°F may require sl;ecial ’ materials for non-metallic components such as oil seals and shims.

9.1.4 Food and drug

The lubricants recommended in this standard are not recommended for food and drug industry

Adjustments to the gear drive selection may be necessary when one or more of the following conditions exists:

- Ambient conditions. Extremes of temperature and environment;

19

ANSIIAGMA 601 O-F97 AMERICAN NATIONAL STANDARD

applications where incidental contact with the product being manufactured occurs.

NOTE: The user must assume the responsibility for selecting the proper lubricant for all food and drug industry applications.

9.15 Mounting position

All gear drives are considered to operate in the manufacturer’s specified mounting position.

9.2 Lubricant viscositv

Lubricant viscosity recommendations are specified by AGMA lubricant numbers. The corresponding viscosity ranges are shown in ANSl/AGMA 9005-D94, table 4 .

expected ambient temperature, and a viscosity which is low enough to allow the oil to flow freely at the start-up temperature but high enough to carry the load at operating temperature.

For synthetic lubricant recommendations, refer to ANSVAGMA 9005D94.

9.4.3 Sump heaters

If a suitable, low temperature gear oil is not available, the gear drive must be provided with a sump heater to bring the oil up to a temperature at which it will circulate freely for starting. The heater should be selected so as to avoid excessive localized heating which could result in rapid degradation of the lubricant.

9.3 Lubrication recommendations 9.5 Lubricant types Recommended lubricants are shown in ANWAGMA 9005D94, table 5.

9.3.1 External cooling

Refer to ANSVAGMA 9005D94, clause 3.

9.5.1 Rust and oxidation inhibited gear lubricants

If the drive lubrication system is equipped with a cooler which limits the oil supply temperature to 125”F, the lubricant grade recommended for 15” to 50°F in ANSI/AGMA 9005D94, table 5 may be also used at the higher temperature range.

9.3.2 Gearing considerations

When there is a large difference in pitch line velocity between the high and low speed gear stages, the use of a lower viscosity lubricant may be more desirable than that recommended in ANWAGMA 9005-D94, table 5. Also a lower viscosity lubricant may be desirable when there is a combination of sleeve and roller element bearings.

9.4 Cold temperature starting

which are more resistant to rust and oxidation than oil without these special features.

These lubricants are commonly referred to as R&O

9.5.2 Anti-scuff (extreme pressure) lubricants

Anti-scuff (extreme pressure (EP)) gear lubricants

gear oils. They are petroleum base liquids which

are petroleum based lubricants containing special chemical

have been formulated to include chemical additives

additives. EP gear lubricants recommended for enclosed gear drives are those containing sulphur, phosphorous or similar type additives. EP gear lubricants should be used only when specified by the gear drive manufacturer (see ANSVAGMA 9005-D94, table 5).

9.4.1 Low temperature conditions

Gear unit lubrication, either by splash or pump, must be given special attention if the unit is to be started or operated at temperatures below which the oil can be effectively splashed or pumped. Preheating the oil may be necessary under these low ambient temperature conditions. The gear manufacturer must always be informed when units are to operate under these conditions.

9.4.2 Low temperature gear oils

Gear drives operating in cold areas must be provided with oil that circulates freely and does not cause high starting torques. An acceptable low temperature gear oil, in addition to meeting AGMA specifications, must have a pour point at least 10°F below the

NOTE: The lead naphthenate type is no longer recommended because of limited availability and poor stability in comperison to the more modern types of lubricants. WARNING: Do not use extreme pressure lubricant or lubricants with formulations including sulphur, chlorine, lead and phosphorous derivatives, as well as graphite and molybdenum disulfide in gear drives equipped with an internal backstop, unless approved by the gear manufacturer or the backstop manufacturer.

9.5.3 Synthetic gear lubricants

Diesters, polyglycols and synthetic hydrocarbons (polyalphaolefins) have been used in enclosed gear drives for special operating conditions. Synthetic lubricants can be advantageous over mineral oils in that they generally are more stable, have a longer life, and operate over a wider temperature range.

20


Synthetics are not “cure-alls”. Each type has different characteristics, and many of them have distinct disadvantages. Such things as compatibility with gear drive and other lube system components, behavior in the presence of moisture, lubricating qualities, overall economics and compatibility with internal coatings should be carefully analyzed for each type of synthetic lubricant under consideration. In the absence of field experience in similar applications, the use of a synthetic lubricant should be carefully coordinated between the user, the gear manufacturer and the lubricant supplier.

CAUTION: Special authorization is required from the manufacturer prior to using a synthetic lubricant in a unit equipped with an internal backstop.

9.5.4 Synthetic lubricant selection

The recommendations for synthetic lubricants are based on gear drive manufacturers’ experience with synthetic hydrocarbons of the polyalphaolefin type. While other types of synthetic lubricants may be used, lack of experience prevents their recommen- dation. The viscosity recommendations may be used as a guide in selection of these other types of lubricants along with the considerations of 9.1.2.

9.6 Maintenance of lubrication system

Lubricants must be free of solid contaminants such as dirt and wear particles, free of water, and contain sufficient additives to maintain their original performance. The best way to maintain lubricant performance is by regular oil changes.

Refer to ANSVAGMA 9005D94, clause 6.

9.6.1 Initial lubricant maintenance

The lubricant in a new gear drive should be drained after 500 hours or four (4) weeks of operation, whichever occurs first. The gear case should be thoroughly cleaned with a commercial grade of flushing oil that is compatible with the seals and operating lubricant.

The original lubricant can be used for refilling if it has been filtered through a filter of 30 microns or less, it is free of water, and the original additive strength is maintained; otherwise, new lubricant must be used. Lubricants should not be filtered through fuller’s earth or any filters which remove lubricant additives.

9.6.2 Subsequent oil change interval

Under normal operating conditions, the lubricant should be changed every 2500 operating hours or

six months, whichever comes first. Conditions that may require more frequent oil change periods include:

- ambient conditions of extreme dust, dirt, moisture and chemical particles or fumes;

- sustained lubricant sump temperatures approaching 200°F;

- duty cycle or ambient conditions causing large and rapid sump temperature changes;

- seasonal ambient temperature changes causing changes in recommended lubricant.

Extending the change period recommended may be preferred based on type of lubricant, amount of lubricant, system down time, or environmental impact of used oil. This can be done through proper implementation of a comprehensive lubricant testing program. As a minimum, the program should include testing for:

- changes in appearance and odor;

- lubricant viscosity (oxidation);

- water concentration;

- contaminant concentration;

- sediment and sludge;

- additive concentration and condition.

In the absence of more specific limits, the guidelines listed as follows may be used to indicate when to change oil:

- water content greater than 0.05% (500 ppm);

- iron content exceeds 150 ppm;

- silicon (dust/dirt) exceeds 25 ppm;

- viscosity changes more than 15%.

These tests should be performed on the initial charge of the gear unit to establish a base line for comparison. Subsequent test intervals should be established based on the unit manufacturer’s and lubricant supplier’s recommendations.

9.6.3 Cleaning and flushing

The lubricant should be drained while the gear drive is at operating temperature. The drive should be cleaned with a flushing oil.

9.6.4 Used lubricants

Used lubricant and flushing oil should be completely removed from the system to avoid contaminating the new charge.

21

ANSI/AGMA 601 O-F97

9.6.5 Cleaning with solvents

The use of a solvent should be avoided unless the gear drive contained deposits of oxidized or contami- nated lubricant which cannot be removed with a flushing oil. When persistent deposits necessitate the use of a solvent, a flushing oil should then be used to remove all traces of solvent from the system.

9.6.6 Inspection

The interior surfaces should be inspected where possible, and all traces of foreign material removed. The new charge of lubricant should be added and circulated to coat all internal parts.

10 Assembly and rotation

10.1 Shaft rotation direction

Rotational direction of both high and low speed shafts is either clockwise (CW) or counterclockwise (CCW). Direction of shaft rotation is determined by

7 L-R

--r L-LR

Plan views

NOTES:

1 LR-L


viewing a specified shaft from a specified free end position. Designation of shaft rotation on drawings or in tables may be shown by letter abbreviations or circular arrows as shown in figure 2.

10.2 Assembly designations

Standard assembly designations are shown in figures 3,4 and 5.

Clockwise rotation

Counter- clockwise

rotation Y

Side End views views

Figure 2 - Shaft rotation

0

BB: R-R

~

R-L

t t

af LR-R 73 R-LR

t

LR-LR

b

Plan views

1. Code: L = Left; R = Right 2. Arrows indicate line of sight to determine direction of shaft extensions. 3. Letters preceding the hyphen refer to number and direction of high speed shaft extensions. 4. Letters following the hyphen refer to number and direction of low speed shaft extensions.

Figure 3 - Parallel shaft spur, helical and herringbone gear drives, single or multiple stage


1-L

Q 1-R

5 t

a 2-L

t

l-LR

0 t

b t

cl+ 2-R

t

r’ n , \ I 2-LR 1

L J

u ’

t Plan views

NOTES:

1. Code: L = Left; R = Right

2. Arrows indicate line of sight to determine direction of shaft extensions.

3. Numerals preceding the hyphen refer to number of high speed shaft extensions.

4. Letters following the hyphen refer to number and direction of low speed shaft extensions.

Figure 4 - Horizontal bevel gear drives, Figure 5 - Vertical bevel gear drives, single stage; horizontal bevel-helical single stage; vertical bevel-helical

drives, multiple stage drives, multiple stage

t views

0 $ 0 1 -UD h+ 2-UD 1 hr- 1

Front views

NOTES: 1. Code: U = Up position-low speed shaft;

D = Down position-low speed shaft.

2. Arrows indicate line of sightto determine direction of shaft extensions.

3. Numerals preceding the hyphen refer to number of high speed shaft extensions.

4. Letters following the hyphen refer to number and direction of low speed shaft extensions.

23

ANSIJAGMA 601 O-F97

11 Ratios

The standard ratios are based upon the (1.5)“.5 geometric numerical progression.

This progression is a modification of the ANSI “10 Series”. The standard ratios are listed in table 11. Exact ratios should be obtained from the manufacturer.

Table 11 - Nominal ratios

Ratio Ratio Ratio Ratio 1.225 9.330 70.62 536.3 1.500 11.39 86.50 658.8 1.837 13.95 105.9 804.5 2.250 17.09 129.7 985.3 2.756 20.93 158.9 1207 3.375 25.63 194.6 1478 4.134 31.39 238.4 1810 5.062 38.44 291.9 2217 6.200 47.08 357.5 2715 7.594 57.67 437.9 3325

12 Marking

12.1 Nameplate data

A suitable nameplate should be attached to the gear drive with the following m inimum information:

- size;

- ratio;

- service power rating;

- high speed shaft rpm;

- service factor;

- lubrication specification.

12.2 AGMA monogram

Use of the AGMA monogram certifies that the manufacturer of the gear drive is a member of AGMA, and does not in any way indicate the manufacturer’s compliance with AGMA standards or practices.


13 Storage

13.1 General

These general storage recommendations should be used when specific manufacturer’s instructions are not available. They apply to gear designs in which the rotating elements are contained in a suitable enclosed housing. See annex E.

Proper protection, storage and inspection of gear drives is considered to be the responsibility of the owner. It is recommended that reducers be stored in a dry, temperature controlled environment. Within this environment, the ambient temperature change should not be allowed to pass through the dew point since this would cause moisture condensation on gear drive surfaces.

13.2 Normal storage

During manufacture and for intervals of storage up to four months, internal components of gear drives should be coated with a suitable oil based rust preventative. This rust preventative should contain water displacement and fingerprint suppressant additives. External machined surfaces should be coated with a similar rust preventative during manufacture. A suitable petroleum base rust preventative should be applied to external surfaces before the drives are placed in storage. Such coating should be self-healing and contain water displacement and fingerprint suppressant additives suitable for protecting the surfaces against rust for a period of up to 12 months.

External and internal inspection of the gear drive should be made monthly. Any moisture observed should be removed at this time and components and surfaces recoated with rust preventative as necessary.

Drives should be inspected prior to charging with lubricant to ensure that no condensate is present in the oil sump.

13.3 Adverse conditions or long term storage

In conditions of long periods of storage or storage in environments subject to high humidity, extreme temperature change, or exposure to an oxidation enhancing atmosphere, gear drives should be completely filled to overflowing with a high quality oil base rust inhibiting lubricant. In cases where it is impractical to fill the gear drive to overflowing, the lubricant should be circulated to coat all internal

24


components and the gear drive sealed to help prevent condensation. Care must be taken to ensure that all enclosed areas are properly vented to prevent the entrapment of moist air.

Ail external machined surfaces should be coated with a solvent based rust inhibiting undercoating and top coated with an asphalt base rust preventative.

Under such adverse or long term storage condition, inspection of the gear drive should be performed on a weekly basis. The sump drain should be opened and a small amount of oil removed along with any condensate which might be present. The drive should be refilled to overflowing and resealed. Outside surfaces should be recoated as necessary.

CAUTION: Some gear drives contain features where overfill with lubricant is not feasible or practical. These features include such items as drywelis on vertical shaft units and labyrinth seals on shaft extensions. For these drives, add the appropriate type and amount of vapor phase rust inhibitor and seal any openings. Inspectthe gear drive on a weekly basis and add the required amount of lubricant. Drives fitted with labyrinth seals cannot be filled with oil or have a vapor phase rust inhibitor installed as both will leak from the unit.

14 Installation

To ensure long service and dependable performance, an enclosed gear drive must be rigidly supported and the shafts accurately aligned. The following describes the minimum precautions required to accomplish this end. The gear reducer manufacturer’s installation manual should be followed, as it may include more detailed procedures than appear in this standard. For owner’s responsibilities, see annex E.

14.1 Foundation

The responsibility for the design and construction of the foundation lies with the user. The foundation must be adequate to withstand normal operating loads and possible overloads which may occur without damage to itself or any of the system components and to maintain alignment of the components under such loads.

14.1 .l Mounting position

Unless a unit is specifically ordered for inclined mounting, the foundation must be level and flat. The lubrication system may not operate properly if the

unit is not mounted in the position for which it is designed. It may be desirable to elevate the foundation to facilitate oil drainage.

14.1.2 Concrete foundation

If a concrete foundation is used, steel mounting pads and bolts of sufficient size to distribute the stress into the concrete should be grouted into the foundation.

14.1.3 Steel foundation

If a structural steel foundation is used (i.e., wide flange beams or channels), a base plate or sole plate of suitable thickness should be used and should extend under the entire unit.

14.2 Foot mounted units

Use shims under the feet of the unit to align the output shaft to the driven equipment. Make sure that all feet are supported so that the housing will not distort when it is bolted down. Improper shimming will reduce the life of the unit and may cause failure. Install dowel pins as instructed by the manufacturer’s installation manual to prevent misalignment and ensure proper realignment if removed for service.

14.3 Shaft mounted units

Shaft mounted drives should be mounted as close to the driven equipment bearing support as possible to minimize bearing loads due to overhung load. Design of the joint connection between the torque arm and the foundation is the user’s responsibility.

14.3.1 Lubricate shafts

Both the hollow shaft and the driven shaft should be liberally lubricated before assembly. The unit must slide freely onto the driven shaft. Do not hammer or force the unit into place.

14.3.2 Axial retention

Follow the manufacturer’s instructions for axial retention of the unit on the driven shaft.

14.3.2.1 Set screws

If set screws are used for axial retention, they should be tightened evenly. Flats may be filed on the driven shaft and a thread locking adhesive used for more positive retention.

14.3.2.2 Thrust plate

In applications which are subject to high vibratory loads, a thrust plate will provide greater resistance to axial movement. Follow the manufacturer’s recommendations for assembly.

25

ANSIjAGMA 601 O-F97 AMERICAN NATIONAL STANDARD

14.3.2.3 Shrink disc

If a shrink disc is used to secure the hollow shaft to the driven shaft, follow the shrink disc manufacturer’s assembly procedure. If the shrink disc manufacturer’s procedures are not available, the following may be used:

a) Any protective spacers between the locking collars should be removed; b) Tighten slightly any three equally spaced locking bolts until the inner ring can just be rotated. Overtightening at this time can damage the inner ring. Measure the gap between the locking collars at several points around the circumference to ensure that they are parallel; c) Slide the shrink disc over the hollow shaft. Tighten the bolts in gradual increments following a circumferential pattern until all bolts are tightened to the specified torque. Check the gap between the locking collars to be sure that they are parallel.

14.3.3 Torque arm

The torque arm should always be mounted within the angular lim its specified by the manufacturer. The preferred mounting position is perpendicular to a line through the output shaft center and the point of attachment of the torque arm to the unit housing. In this position the m inimum load on the torque arm will be experienced. Design of the joint connection between the torque arm and foundation is the user’s responsibility.

14.4 Prime mover mounting

Align the prime mover to the unit input shaft using shims under the feet. Make sure that the feet are supported. Dowel the prime mover to its foundation.

14.5 Shaft connections

14.5.1 Fits

Clearance or interference fits for coupling hubs should be in accordance with ANSVAGMA QOO2-A86. Outboard pinion and sprocket fits should be as recommended by the pinion or sprocket manufacturer. Coupling hubs, pinions and sprockets with interference fits should be heated according to the manufacturer’s recommendations, generally 250°F to 3OO”F, before assembling to the shaft.

14.5.2 Location

Coupling hubs should be mounted flush with the shaft ends, unless specifically ordered for overhung

mounting. Pinions, sprockets and sheaves should be mounted as close as possible to the unit housing to m inimize bearing loads and shaft deflection.

14.5.3 Coupling alignment

Shaft couplings should be installed according to the gear manufacturer’s recommendations for gap, angular and parallel alignment. In many installa- tions, it is necessary to allow for thermal and mechanical shaft movement when determining shaft alignment. The gear manufacturer’s recommendations should be followed. Where no recommendations are made by the gear drive manufacturer, follow the coupling manufacturer’s recommendations.

14.5.3.1 Axial displacement

The gap between shaft ends should be the same as the specified coupling gap unless overhung mounting of the coupling hub is specified. The coupling gap and shaft gap must be sufficient to accommodate any anticipated thermal or mechanical axial movement.

14.5.3.2 Angular alignment

Insert a spacer or shim stock equal to the required coupling gap between the coupling hub faces and measure the clearance using feeler gauges. Repeat this at the same depth at 90 degree intervals to determine the amount of angular m isalignment.

14.5.3.3 Parallel alignment

Mount a dial indicator to one coupling hub, and rotate this hub, sweeping the outside diameter of the other hub. The parallel m isalignment is equal to one-half of the total indicator reading. Another method is to rest a straight edge squarely on the outside diameter of the hubs at 90 degree intervals and measure any gaps with feeler gauges. The maximum gap measurement is the parallel m isalignment.

14.5.3.4 Checking alignment

After both angular and parallel alignments are within specified lim its, tighten all foundation bolts securely and repeat the above procedure to check alignment. If any of the specified lim its for alignment are exceeded, realign the coupling.

14.5.4 Sprocket or sheave alignment

Align the sheaves or sprockets square and parallel by placing a straight edge across their faces. Alignment of bushed sheaves and sprockets should be checked after bushings have been tightened.

26

Check horizontal shaft alignment by placing a level vertically against the face of the sheave or sprocket. Adjust belt or chain tension per the manufacturer’s specified procedure.

14.5.5 Outboard pinion alignment

Align the pinion by adjusting the gear tooth clearance according to the manufacturer’s recommendations and checking for acceptable outboard pinion tooth

contact. The foundation bolts may have to be loosened and the unit moved slightly to obtain this contact. When the unit is moved to correct tooth contact, the prime mover should be realigned.

14.5.6 Rechecb alignment

After a period of operation, recheck alignment and adjust as required.

27

ANSIIAGMA 601 O-F97

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28


Annex A (informative)

Service factors

rhe foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSIIAGMA 6010-F97, Standard for Spur Helical, Herringbone and Bevel Enclosed Drives.]

A.1 Purpose

This annex provides a detailed guide for determining service factors for enclosed gear drives.

A.1 .l Selection of service factors

Before an enclosed speed reducer or increaser can be selected for any application, an equivalent unit power rating (service factor = 1 .O) must be determined. This is done by multiplying the specified power by the service factor. Since the service factor represents the normal relationship between the gear unit rating and the required application power, it is suggested that the service factor be applied to the nameplate rating of the prime mover or driven machine rating, as applicable.

Manufacturer and user must agree upon which power, prime mover rating or driven machine requirements, should dictate the selection of the gear drive. It is necessary that the gear drive selected have a rated unit capacity equal to or in excess of this “equivalent unit power rating”.

All service factors listed are 1 .O or greater. Service factors less than 1 .O can be used in some applications when specified by the user and agreed to by the manufacturer.

Table A.2 should be used with caution, since much higher values have occurred in some applications. Values as high as ten have been used. On some applications up to six times nominal torque can occur, such as: Turbine/Generator drives, Heavy Plate and Billet rolling mills.

A.2 Listing of service factors

The table of service factors has been developed from the experience of manufacturers and users of gear drives for use in common applications. It is suggested that service factors for special

applications be agreed upon by the user and the gear manufacturer when variations of the values in the table may be required.

A.3 Determining service factors

In addition to the tables, an analytical approach may be used to determine the service factor. See 8.3 for the important factors to be considered.

A.4 Service factor tables

Service factors have served industry well when the application has been identified by knowledgeable and experienced gear design engineers. The tables are provided for information purposes only and should be used only after taking into account all of the external influences which may affect the operation of the enclosed gear drive.

A.4.1 Use of tables

Service factors shown in table A.2 are for gear drives driven by motors (electric or hydraulic) and turbines (steam or gas).

A.4.2 Driver influence

When the driver is a single cylinder or multi-cylinder engine, the service factors from table A.2 must be converted to the values from table A.1 for the appropriate type of prime mover.

A.5 Example

If the application is a centrifugal blower, the service factor from table A.2 is 1.25 for a motor or turbine. Table A.1 converts this value to 1.50 for a multi- cylinder engine and 1.75 for asingle cylinder engine.

CAUTION: Any user of enclosed gear drives should make sure he has the latest available data on the factors affecting the selection of a gear drive. When better load intensity information is available on the driving or driven equipment, this should be considered when a service factor is selected.

29

ANSIIAGMA WlO-F97 AMERICAN NATIONAL STANDARD

Table A.1 - Conversion table for single or multi-cylinder engines to find equivalent single or multi-cylinder service factor

Steam and gas turbines, hydraulic or electric motor Single cylinder engines Multi-cylinder engines

I .oo 1.50 1.25 1.25 1.75 1.50 1.50 2.00 I.75 I .75 2.25 2.00 2.00 2.50 2.25 2.25 2.75 2.50 2.50 3.00 2.75 2.75 3.25 3.00 3.00 3.50 3.25

Table A.2 - Service factors for enclosed gear drives driven by motors (hydraulic or electric) or turbines (steam or gas)

Application

Agitators (mixers) Pure liquids Liquids and solids Liquids - variable density

Blowers Centrifugal Lobe Vane

Brewing and distilling Bottling machinery Brew kettles - continuous duty Cookers - continuous duty Mash tubs - continuous duty Scale hopper - frequent starts

Can filling machines Car dumpers Car pullers Clarifiers Zlassifiers

Clay working machinery Brick press Briquette machine Pug m ill

Zompactors >ompressors

Centrifugal Lobe Reciprocating, multi-cylinder Reciprocating, single-cylinder

T Load duration 3 to IO hours

per day Up to 3 hours

per day

I .oo 1 .oo 1 .oo

I .oo 1 .oo I .oo

1.00 I .25 I.25 1.25 I .25 I .oo 1.50 I .oo 1.00 I .oo

I .50 I .50 1 .oo 2.00

1.00 1 .oo 1.50 1.75

Dver IO hours per day

I .oo 1.25 I .25

I .25 I .50 I .50

I .oo I .25 I.25

I.25 I .50 1.50

1 .oo I .25 I .25 1.25 1.25 1 .oo I .75 1.25 1 .oo I.25

I.25 1.25 I .25 I .25 I .50 1.25 2.00 I .50 I.25 I .50

1.75 I.75 I.25 2.00

2.00 2.00 I .50 2.00

I .oo I .25 I .25 I .50 I .50 I.75 I.75 2.00

(continued)

30

AMERICAN NATIONAL STANDARD ANSIJAGMA 6010-F97

Table A.2 (continued)

Application

Cranes’) Dry dock

Main hoist Auxiliary hoist Boom hoist Slewing drive Traction drive

Container Main hoist Boom hoist Trolley drive

Gantry drive Traction drive

M ill duty Main hoist Auxiliary Bridge Trolley travel

Industrial duty Main Auxiliary Bridge Trolley travel

Crusher Stone or ore

Dredges Cable reels Conveyors Cutter head drives Pumps Screen drives Stackers Winches

Elevators Bucket Centrifugal discharge Escalators Freight Gravity discharge

Extruders General Plastics

Variable speed drive Fixed speed drive

Rubber Continuous screw operation Intermittent screw operation

Load duration Up to 3 hours 3 to 10 hours Over 10 hours

per day per day per day

2.50 2.50 2.50 2.50 2.50 3.00 2.50 2.50 3.00 2.50 2.50 3.00 3.00 3.00 3.00

3.00 3.00 3.00 2.00 2.00 2.00

3.00 3.00 3.00 2.00 2.00 2.00

3.50 3.50 3.50 3.50 3.50 3.50 2.50 3.00 3.00 2.50 3.00 3.00

2.50 2.50 3.00 2.50 2.50 3.00 2.50 3.00 3.00 2.50 3.00 3.00

1.75 1.75 2.00

1.25 1.25 1.50 1.25 1.25 1.50 2.00 2.00 2.00 2.00 2.00 2.00 1.75 1.75 2.00 1.25 1.25 1.50 1.25 1.25 1.50

1 .oo 1.25 1.50 1 .oo 1.00 1.25 1 .oo 1.00 1.25 1.00 1.25 1.50 1 .oo 1.00 1.25

1.50 1.50 1.50

1.50 1.50 1.50 1.75 1.75 1.75

1.75 1.75 1.75 1.75 1.75 1.75

(continued)

31

ANSIJAGMA 601 O-F97 AMERICAN NATIONAL STANDARD

Application

=ans Centrifugal Cooling towers Forced draft Induced draft Industrial & m ine

=eeders Apron Belt Disc Reciprocating Screw

=ood industry Cereal cooker Dough m ixer Meat grinders Slicers

Senerators and exciters iammer m ills -foists

Heavy duty Medium duty Skip hoist

-sundry Tumblers Washers

-umber industry Barkers - spindle feed

Main drive Conveyors - burner

Main or heavy duty Main log Re-saw, merry-go-round

Conveyors Slab Transfer

Chains Floor Green

Cut-off saws Chain Drag

Debarking drums Feeds

Edger Gang Trimmer

Table A-2 (continued)

Load duration Up to 3 hours 3 to 10 hours Over 10 hours


1.00 1.00 1.25 2.00 2.00 2.00 1.25 1.25 1.25 1.50 1.50 1.50 1.50 1.50 1.50

1 .oo 1.25 1.50 1 .oo 1.25 1.50 1 .oo 1 .oo 1.25 1.50 1.75 2.00 1 .oo 1.25 1.50

1.00 1 .oo 1.25 1.25 1.25 1.50 1.25 1.25 1.50 1.25 1.25 1.50 1 .oo 1.00 1.25 1.75 1.75 2.00

1.75 1.75 2.00 1.25 1.25 1.50 1.25 1.25 1.50

1.25 1.25 1.50 1.50 1.50 2.00

1.25 1.25 1.50 1.75 1.75 1.75 1.25 1.25 1.50 1.50 1.50 1.50 1.75 1.75 2.00 1.25 1.25 1.50

1.75 1.75 2.00 1.25 1.25 1.50

1.50 1.50 1.50 1.50 1.50 1.75

1.50 1.50 1.75 1.50 1.50 1.75 1.75 1.75 2.00

1.25 1.25 1.50 1.75 1.75 1.75 1.25 1.25 1.50

,---r:-* .^-I,

32



Application

Lumber industry (continued) Log deck Log hauls - incline - well type Log turning devices Planer feed Planer tilting hoists Rolls - live-off bearing - roll cases Sorting table Tipple hoist Transfers

Chain Craneway

Tray drives Veneer lathe drives

Metal mills Draw bench carriage and main drive Runout table

Non-reversing Group drives Individual drives

Reversing Slab pushers Shears Wire drawing Wire winding machine

Metal strip processing machinery Bridles Coilers & uncoilers Edge trimmers Flatteners Loopers (accumulators) Pinch rolls Scrap choppers Shears Slitters

Mills, rotary type Ball & rod Spur ring gear Helical ring gear Direct connected Cement kilns Dryers & coolers

Mixers Concrete

‘aper mills2) Agitator (mixer) Agitator for pure liquors Barking drums

I Load duration Up to 3 hours 3 to 10 hours 3ver 10 hours


1.75 1.75 1.75 1.75 1.75 1.75 1.75 1.75 1.75 1.25 1.25 1.50 1.50 1.50 1.50 1.75 1.75 1.75 1.25 1.25 1.50 1.25 1.25 1.50

1.50 1.50 1.75 1.50 1.50 1.75 1.25 1.25 1.50 1.25 1.25 1.50

1.25 1.25 1.50

1.50 1.50 1.50 2.00 2.00 2.00 2.00 2.00 2.00 1.50 1.50 1.50 2.00 2.00 2.00 1.25 1.25 1.50 1.25 1.50 1.50

1.25 1.25 1.50 1.00 1 .oo 1.25 1.00 1.25 1.50 1.25 1.25 1.50 1.00 1 .oo 1.25 1.25 1.25 1.50 1.25 1.25 1.50 2.00 2.00 2.00 1 .oo 1.25 1.50

2.00 1.50 2.00 1.50 1.50

2.00 1.50 2.00 1.50 1.50

2.00 1.50 2.00 1.50 1.50

1.25 1.25 1.50

1.50 1.50 1.25 1.25 2.00 2.00

1.50 1.25 2.00 (continued,

1

33

ANSIJAGMA 6010-F97 AMERICAN NATIONAL STANDARD


Application

Paper m ills?) (continued) Barkers - mechanical Beater Breaker stack Calender31 Chipper Chip feeder Coating rolls Conveyors

Chip, bark, chemical Log (including slab)

Couch rolls Cutter Cylinder molds Dryers31

Paper machine Conveyor type

Embosser Extruder Fourdrinier rolls (includes lump breaker, dandy roll,

wire turning, and return rolls) Jordan Kiln drive M t. Hope roll Paper rolls Platter Presses - fett & suction Pulper Pumps - vacuum Reel (surface type) Screens

Chip Rotary Vibrating

Size press Super calendefi) Thickener (AC motor)

(DC motor) Washer (AC motor)

(DC motor) Wind and unwind stand Winders (surface type) Yankee dryers31

Plastics industry Primary processing

Intensive internal m ixers Batch m ixers Continuous m ixers

T Load duration 3 to 10 hours

per day Up to 3 hours

per day Dver 10 hours

per day

2.00 2.00 2.00 1.50 1.50 1.50 1.25 1.25 1.25 1.25 1.25 1.25 2.00 2.00 2.00 1.50 1.50 1.50 1.25 1.25 1.25

1.25 1.25 1.25 2.00 2.00 2.00 1.25 1.25 1.25 2.00 2.00 2.00 1.25 1.25 1.25

1.25 1.25 1.25 1.25 1.25 1.25 1.25 1.25 1.25 1.50 1.50 1.50 1.25 1.25 1.25 1.50 1.50 1.50 1.50 1.50 1.50 1.25 1.25 1.25 1.25 1.25 1.25 1.50 1.50 1.50 1.25 1.25 1.25 2.00 2.00 2.00 1.50 1.50 1.50 1.25 1.25 1.25

1.50 1.50 1.50 1.50 1.50 1.50 2.00 2.00 2.00 1.25 1.25 1.25 1.25 1.25 1.25 1.50 1.50 1.50 1.25 1.25 1.25 1.50 1.50 1.50 1.25 1.25 1.25 1 .oo 1 .oo 1 .oo 1.25 1.25 1.25 1.25 1.25 1.25

1.75 1.50

1.75 1.50

1.75 1.50

(continued,

1

34



Load duration Application Up to 3 hours 3 to 10 hours Over 10 hours

per day per day per day Plastics industry (continued)

Batch drop mill - 2 smooth rolls 1.25 1.25 1.25 Continuous feed, holding & blend mill 1.25 1.25 1.25 Compounding mill 1.25 1.25 1.25 Calenders 1.50 1.50 1.50

Secondary processing Blow molders 1.50 1.50 1.50 Coating 1.25 1.25 1.25 Film 1.25 1.25 1.25 Pipe 1.25 1.25 1.25 Pre-plasticizers 1.50 1.50 1.50 Rods 1.25 1.25 1.25 Sheet 1.25 1.25 1.25 Tubing 1.25 1.25 1.50

%rllers - barge haul 1.25 1.25 1.50

Pumps Centrifugal 1 .oo 1 .oo 1.25 Proportioning 1.25 1.25 1.50 Reciprocating

Single acting, 3 or more cylinders 1.25 1.25 1.50 Double acting, 2 or more cylinders 1.25 1.25 1.50

Rotary Gear type 1 .oo 1 .oo 1.25 Lobe 1 .oo 1 .oo 1.25 Vane 1 .oo 1 .oo 1.25

qubber industry Intensive internal mixers

Batch mixers 1.75 1.75 1.75 Continuous mixers 1.50 1.50 1.50

Mixing mill - 2 smooth rolls (if corrugated rolls are used, then use the same service factors that 1.50 1.50 1.50 are used for a cracker warmer)

Batch drop mill - 2 smooth rolls 1.50 1.50 1.50 Cracker warmer - 2 rolls; 1 corrugated roll 1.75 1.75 1.75 Cracker - 2 corrugated rolls 2.00 2.00 2.00 Holding, feed & blend mill - 2 rolls 1.25 1.25 1.25 Refiner - 2 rolls 1.50 1.50 1.50 Calenders 1.50 1.50 1.50

sand muller 1.25 1.25 1.50

sewage disposal equipment Bar screens 1.25 1.25 1.25 Chemical feeders 1.25 1.25 1.25 Dewatering screens 1.50 1.50 1.50 Scum breakers 1.50 1.50 1.50 Slow or rapid mixers 1.50 1.50 1.50 Sludge collectors 1.25 1.25 1.25 Thickeners 1.50 1.50 1.50 Vacuum filters 1.50 1.50 1.50

(continued)

35


Table A-2 (concluded)

Load duration Application Up to 3 hours 3 to 10 hours Over 10 hours

per day per day per day Screens

Air washing 1 .oo 1 .oo 1.25 Rotary - stone or gravel 1.25 1.25 1 so Traveling water intake 1.00 1 .oo 1.25

Sugar industry Beet slicer 2.00 2.00 2.00 Cane knives 1.50 1.50 1.50 Crushers 1.50 1.50 1.50 Mills (low speed end) 1.75 1.75 1.75

Textile industry Batchers 1.25 1.25 1.50 Calenders 1.25 1.25 1.50 Cards 1.25 1.25 1 so Dry cans 1.25 1.25 1.50 Dryers 1.25 1.25 1.50 Dyeing machinery 1.25 1.25 1.50 Looms 1.25 1.25 1.50 Mangles 1.25 1.25 1.50 Nappers 1.25 1.25 1.50 Pads 1.25 1.25 1.50 Slashers 1.25 1.25 1.50 Soapers 1.25 1.25 1.50 Spinners 1.25 1.25 1.50 Tenter frames 1.25 1.25 1.50 Washers 1.25 1.25 1.50 Winders 1.25 1.25 1.50

UOTES: I) Crane drives are to be selected based on gear tooth bending strength. Contact gear manufacturer for strength .atings. Service factor in durability should be a minimum of 1 .O. 3 Service factors for paper mill applications are applied to the nameplate rating of the electric drive motor at the notor rated based speed. $1 Anti-friction bearings only. Use 1.5 for sleeve bearings. 0 A service factor of 1 .OO may be applied at base speed of a super calender operating over-speed range of part ange constant horsepower, part range constant torque where the constant horsepower speed range is greater than I .5 to 1. A service factor of 1.25 is applicable to super calenders operating over the entire speed range at constant orque or where the constant horsepower speed range is less than 7.5 to 1.

36


Annex B (informative)

Keys and keyways for shaft extensions

phe foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANWAGMA 6010-F97, Standard for Spur, Helical, Herringbone and Bevel Enclosed Drives.]

B.l Purpose

The purpose of this annex is to present recommended standard sizes of keys and keyways to be used on shaft extensions of standard gear reducers. This annex is in conformance with ANSI B17.1-1967 and ANSI/AGMA 9002-A86.

The intent of the recommended standardization is to simplify the accommodation of power transmission accessories on these gear drives. These accessories typically include shaft couplings, sheaves and sprockets.

While it is recognized that there are occasional reasons to deviate from these recommendations, general conformance will eliminate confusion and m isinterpretation between vendor or supplier and user as well as ensure compatibility.

B.2 Keys

Recommended key sizes are shown in table B.l. Square keys are preferred through 6.5 inches shaft diameter. Rectangular keys are preferred over 6.5 inches shaft diameter.

For tapered shafts, the largest tapered diameter determines the key size.

B.3 Keyways

Keyway sizes should normally be selected to result in a commercial fit with the key. A commercial fit will result in a clearance fit with the sides of a key.

Under certain circumstances, it may be necessary to provide a radius in the keyway. The recommendations for keyway radii and key chamfer are shown in table 8.2.

Table B.l - Key sizes and tolerances for square and rectangular keys (inches)

Nominal shaft Recommended key size Recommended key width & height tolerance l) diameter Commercial class Precision class

Over To Square Rectangular Square Rectangular Square Rectangular (Incl.) Width & hgt. Width & hgt. Width & hgt. Width Height

0.3125 0.4375 O.OS37xO.0937 - - +o.ooo/-0.002 +o.ooo/-0.003 +O.OOl/-0.000 +O.OOl/-0.000 +0.005/-0.005

0.4375 0.5625 0.1260x0.1250 0.1250~0.0937 +0.000/-0.002 +O.OOO/-0.003 +0.001/-O 000 +O.OOl/-0.000 +0.005/-O 005

0.5625 0.6750 0.1875x0.1875 0.1875x0.1250 +O.OOO/-0.002 +O.OOO/-0.003 +O.OOl/-0.000 +O.OOl/-0.000 +0.005/-0.005

0.8750 1.2500 0.2500x0.2500 0.2500x0.1875 +O.OOO/-0.002 +O.OOO/-0.003 +O.OOl/-0.000 +O.OOl/-0.000 +0.005/-0.005

1.2500 1.3750 0.3125x0.3125 0.3125x0.2500 tO.OOO/-0.002 tO.OOO/-0.003 tO.OOl/-0.000 tO.OOl/-0.000 +0.005/-0.005

1.3750 1.7500 0.3750x0.3750 0.3750x0.2600 to.ooo/-0.002 to.ooo/-0.003 tO.OOl/-0.000 tO.OOl/-0.000 +0.005/-0.005

1.7500 2.2500 0.6000x0.5000 0.5ooox0.3750 to.ooo/-0.002 to.ooo/-0.003 tO.OOl/-0.000 +O.OOl/-0.000 +0.005/-0.005

2.2500 2.7500 0.8250~06250 0.6250x0.4375 tO.OOO/-0.002 tO.OOO/-0.003 tO.OOl/-0.000 tO.OOl/-0.000 +0.005/-0.005

2.7500 3.2500 0.7600x0.7500 0.7500~0.5000 tO.OOO/-0.002 tO.OOO/-0.003 tO.OOl/-0.000 tO.OOl/-0.000 +0.005/-0.006

3.2500 3.7500 0.8750x0.8750 0.8750x0.6250 tO.OOO/-0.003 tO.OOO/-0.004 tO.OOl/-0.000 tO.OOl/-0.000 +0.005/-0.005

3.7500 4.5000 1.0000x1.0000 1.coOOx0.7500 to.ooo/-0.003 to.ooo/-0.004 tO.OOl/-0.000 tO.OOl/-0.000 +0.005/-0.005

4.5000 5.5000 1.2500x1.2500 1.25QOxO.875O tO.OOO/-0.003 tO.OOO/-0.004 tO.OOl/-0.000 tO.OOl/-0.000 +0.005/-0.005

5.5000 6.5000 1.5CQOx1.5000 1.5000x1.0000 tO.OOO/-0.003 tO.OOO/-0.004 +0.002/-0.000 +0.002/-0.000 +0.005/-0.005

6.5000 7.5000 1.7500x1.7500 1.7500~15000 tO.OOO/-0.004 tO.OOO/-0.005 +0.002/-0.000 +0.002/-0.000 +0.005/-0.005

7.5000 9.0000 2.0000x2.0000 2.mx1.5000 to.ooo/-0.004 to.ooo/-0.005 +0.002/-0.000 +0.002/-0.000 +0.005/-0.005

NOTE: l) Tolerances agree with ANSI B17.1-1967, Reaffirmed 1989.

37

ANSIIAGMA 601 O-F97

Table B.2 - Values for keyway fillet radius and suggested key chamfer

38


Annex C (informative)

Illustrative examples

rhe foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSVAGMA 6010-F97, Standard for Spur, Helical, Herringbone and Bevel Enclosed Drives.]

Introduction

The examples shown in this annex are based on the assumption that the gear set is the minimum rated component. In practice all component ratings must be calculated to determine the lowest rated components. Tables, figures and equation references are taken from ANSVAGMA 2001-C95 and ANSI/ AGMA 2003-A86.

C.l Uniform loading

Calculate the application horsepower for the gears of a single reduction helical speed reducer used to drive a uniformly loaded conveyor belt. The input shaft and pinion are driven by an 870 rpm motor. The gear set is represented by the following data:

Item Pinion Gear Number of teeth 31 93 Diametral pitch, normal 8 Pressure angle, normal 20” Helix angle 15” Face width, inch 2.00 Material Grade 1

Carburized Steel Hardness 60 HRC Center distance, inch 8.0015

Both gear and pinion are standard addendum, cut with standard pre-shave hobs, and shaved. Heat treat distortion is controlled to produce both gears as AGMA Quality Level 8.

The allowable horsepower of the gear set will be determined for both surface pitting strength and root bending strength. The application horsepower, PaI of the gear set will be the lesser of these allowable horsepowers.

Surface pitting allowable power at unity service factor:

P acu

(see 2001, Eq. 27)

np = 870 rpm F = 2.00 in I = 0.192

d - 2(8*0015) - 4 001 in 3+1 *

x, = 1.25 (see 2001, Eq. 23; vf = 911 ft/min and 12,=8)

K, = 1.30 (see 2001, Eq. 36)

CP = 2300 [lb/inq0.5

kc = 180 000 lb/in2 (see 2001, table 3) N = 870 cycles/min (60 min/hr) (5000 hr)

= 2.61 x lo8 cycles

ZN - = 1 4488 (2.61 x 108)-o.023 = 0.9277 (see 2001, figure 17)

CH = 1 .O since gears are surface hardened

P acu = 87o(2.00) 0.192 126 000 (1.30)(1.25) 2 4.001(180 000)(0.9277)

2300 1 = 138 hp

Bending allowable power at unity service factor for the pinion is calculated as follows:

P n,d atu l FJsY

=m~K,K, d se: 2001, Eq. 28)

K, = 1.25

K, = 1.30

JP = 0.5226

pd = 8 cos 15” = 7.7274

4s = 55 000 lb/in2 (see 2001, table 4)

YN = 1.3558 (2.61 x 1 08)-o.0178 = 0.9602 (see 2001, figure 18)

P 870(4.001)

an4 = 126000 (1.241.30) . 2.0;~g426)(55 000)

x (0.9602) = 121 hp Bending allowable power at unity service factor for the gear is calculated as follows:

JG = 0.5664

N = 2.61 x 1 OS cycles

’ = 8.7 x 10’ cycles

39

ANSI/AGMA 6010-F97

YN = 1.3558 (8.7x 1 07)-o.o’78 = 0.9792

p =870(4.001) ahl 126000 (1.25;jl.30) .

2*o;F$464) (55 000)

x (0.9792) = 134 hp

This gearset is lim ited by the root bending allowable power of the pinion, 121 horsepower. To include the 1 .10 safety factor in bending (see 5.2)) increase the service factor by 10%. Therefore:

k& = 1.25 x 1 .10 = 1.375 (see 2001, annex A, table A.2)

pa - 1.375 - 121 = 88 hp

C.2 Variable loading

It is desired to use a gearset of existing design in a new application which requires a life of 2000 hours. This gear set will be used in an installation which has an overload factor of 1 .O by agreement between the user and the manufacturer. The gear set will be subjected to the multi-load cycle of:

Percent time used divided by

100, .q 0.90

0.05

0.05

Since this is an existing gearset, the following gear data is given:

Item Number of teeth Diametral pitch, normal

~ Pressure angle, normal Helix angle Outside diameter, inch Face width, inch Material

I Hardness Center distance, inch

Input power,

P 18.2 hp

38.7 hp

54.0 hp

Pinion speed, JQ 712 rpm = 42 720 rph 712 rpm = 42 720 rph 712 rpm = 42 720 roh

Pinion Gear 15 66

8 20” 20”

2.350 9.135 1.750 1.750

Grade 2 Carburized steel

60 HRC 5.5015

Both gears are AGMA quality level 8.

This gear set must be analyzed for both surface pitting strength and bending strength. The pitting strength will be checked first by calculating the contact stress value for each of the three duty cycle

parts. Referring to figure 17 of ANSI/AGMA 2001 -C95, the life cycles for each of the three stress levels can be found. These life cycles must be combined using M iner’s Rule to determine the total life hours of the pinion and gear for pitting resistance.

SC = cp J W&K& Kn =f do 7

(see 2001, Eq. 1)

c, = 2300 [lb/in2]0.5 & = 1.0

& = 1.14 (Q,,=O and vt = 380 fpm) & = 1.0 &I = 1.25 Cf = 1.0

d = 2(5.5015) (&)=2.038 in I = 0.1995 w t = 126 OOW

nPd

w il

= 126 OOO(18.2) 712(2.038) = 1580 Ibs

w t2

= 126000(38.7) 712(2.038) = 3360 Ibs

w t3

= 126000(54.0) 712(2.038)

= 4689 Ibs

SC1 = 2300 J 1580(1.25)(1.14)

(2.038)(1.75)(0.1995) = 12' 381 Ibfin2

SC2 = 2300 J 3360(1.25)(1.14) (2.038)(1.75)(0.1995) = lg8 674'bh2

SC3 = 2300 J

4689(1.25)(1.14) = (2.038)(1.75)(0.1995) 222 886 IbIn

As this is grade 2 carburized steel, sac = 225 000 lb/in2 (see 2001, table 3).

‘Nl =129=O575O

225000 * 2, = ;z ;; = 0.8386

4v3 = 222 = 0 - 9906 225000

Solving the equation 2, = 1.4488 Wo.o23 for N;

N, =

N 2

N 3

= 2.87 x 1017cycles

= 1.52 x 107cycles

40


The total number of hours of pitting resistance life for the pinion is:

Y Lrl .I”& = w = 0.7566 b3

635 = 1.0559 IId 03 000 Solving the equation YN = 1.3558 N-0.017* for N

= 7111 hours

N 1 x 1032cycles

= 1.73 x 1014 cycles

As YN3 > 1 .O, solve the equation YN = 6.151 4N-0.11g2 The total number of hours of pitting resistance life for the gear is:

Life=[(m)+(m)

+ (m)r = 31311hours

Both gears will exceed the required 2000 hours life in pitting resistance. They must next be checked for bending strength by calculating the bending stress for each of the three duty cycle parts. The life cycles for each of the three stress levels can be found. These life cycles must be combined using M iner’s Rule to determine the total life hours of the pinion and gear for root bending strength.

St = W&&K, F 5 !&f!% (see 2001, Eq. 10) J

K, = 1.0

& = 1.14

pd = 8 cos 20” = 7.5175 in-l x, =l.O K, =1.25

JP = 0.4182

JG = 0.4360 8& =l.O

The stresses and total number of bending strength life hours for the pinion are:

$1 = 1580(7.5175)(1.25)(1.14) = (1.75)(0.4182) 23 127 lb/i,,2

St2 = 3;60(7.5175)(1.25)(1.14) = 4g 182 lb/i,,2

(1.75)(0.4182)

St3 = 4689(7.5175)(1.25)(1.14) = 68 635 lb/in2

(1.75)(0.4182)

for N;

= 3.96 x lo6 cycles

Life=[(m)+(w)

+ (w)r = 1854hours

As can be seen, this gear set will not reach the 2000 hours life requirement, as the pinion teeth will theoretically fracture at 1854 hours.

C.3 Overload conditions

A pinion stand drive of the characteristics shown is expected to be subjected to infrequent (less than 100) momentary overloads. Determine the maximum peak momentary overload to which the gear set may be subjected without the teeth yielding.

say ‘d Kmy KY 2 W,,,,-- F JK,t (see 2001, Eq. 45)

Gear Set Data:

I? = 1.6 pd = 3 in-’

F = 9.0 in d = 9.375 in JP = 0.517

Material: Steel, 340 HB m in, therefore:

say = 131 080 lb/in2 (see 2001, figure 16) I$ = 0.75 (industrial practice) K& =0.0144(g) + 1.07 = 1.1996 (see 2001, Eq. 46)

Solving Eq. 45 for Wmax:

(131080)(0.75) 1 Wmm($) (o(:;‘;$,

203 373 lb 2 Wi, As this is grade 2 carburized steel, Q = 66 000 lb/in2 The maximum momentary peak overload allowable (see 2001, table 4). is 203 373 Ibs. Converting this load to torque yields:

‘Nl = M = 0.3558 Wt d (203 372)(9.375) T=2= 2 = 953 310 lb in

41

item Pinion Gear Number of teeth 18 59 Diametral pitch 2.1 Pitch diameter, inch 8.571 28.095 Face width, inch 3.75 Pressure angle 20" Spiral angle 35" Mean cone distance, inch 12.811 Cutter radius, inch 9.0 Tooth taper Standard Tool edge radius 12/pd Hardness (C&H) min. 55 HRC 55 HRC Desian life 10 000 hrs

The application is considered to be “general commercial”. The gearset is manufactured to AGMA quality 8 tolerances and is lapped. The gears are crowned. One member is straddle mounted.

ANSIIAGMA 601 O-F97

C.4 Uniform loading

Calculate the application horsepower for a spiral bevel gear set, with the following data, used in a single reduction right angle reducer used to drive a uniform load at 870 rpm input, driving a uniform load:

The surface pitting allowable power at unity service factor is calculated as follows:

. ..(see 2003, Eq. 5.4)

!P = 870 rpm

vt =1950 fpm I = 0.132 (see 2003, appendix C) C,=Cf=C,=CH=CT=CR=l.o

C mf = 1.32 (see 2003, table 2) C xc = 1.5 (see 2003, clause 11) S ac = 180 000 lb/in*

cp = 2290 (lb/in2)0.5

cb =0.735

8 - --55000 ' - 20.5(S) 125

30x106+30x106 = 0.385

. ..(see 2003, Eq. 8.4)

K, = 85- lO(O.385) = 81.146 . ..(see 2003, Eq. 8.2)

0.385

c, = 81.146 81.146 + @%

= 0.846 . ..(see 2003, Eq. 8.1)

Cm=Km= [1.2+(F)](l.32)-1.584

NOTE: TV = T- . ..(see 2003, Eq. 10.1)

p,, - 870 (3.75) 0.132(0.846) 126 000 1(1.584)(1)(1)(1.45)

2 180 OOO(8.571) 0.911(l)

2290(0.735) l(l) = 848 hp

Bending allowable power at unity service factor for the pinion is calculated as follows:

nPF JK,K,, Satd KL Pat = -

126 000 KS Km Ka p, G

. . . (see 2003, Eq. 5.8) J = 0.271 (see 2003, appendix C) K;=c,=O.846

k&=&=1.584

S

it

= 55 000 lb/in* =0.88 (see 2003, figure 2)

&=K~=&=1.0 0.279

4 = log&ill35") = -1.156 . ..(see 2003, clause 12)

-1.156 Kx = k211 + 0.789 = 1.107

. ..(see 2003, Eq. 12.1)

- K _ s- P-8 2-l”.25) = l 330 1.2

pat = 870(3.75) 0.271 (1.107)(0.846)

126000 1.330 (1.584)(l)

x 55000(8.571) - 0.880 = 616 hp 2.1 l(l)

C.5 Thermal example

Calculate the thermal rating of a two stage helical gear reducer operating at 1780 rpm input in a counterclockwise direction of rotation. The ambient air temperature is a maximum of 100°F and the reducer sump temperature shall be limited to 200°F. The unit will be operated in a large indoor space, located at sea level, where the air velocity will exceed 275 FPM. The gear reducer is splash lubricated with IS0 220 weight mineral oil, and is subjected to continuous operation. The starting point depends upon experience. The final iterative solution is shown for this example.

The solution for the thermal rating of a gear reducer involves balancing the heat generated under operating conditions with the heat dissipation capability of

42


the gear case. Since the heat generated under operating conditions consists of both load dependent and non-load dependent losses, the solution becomes iterative.

The final iterative value for this example is based on 105 hp.

C.5.1 Calculate heat generation

C.5.1.1 Bearing power losses, Pw and I’J.VB

The reducer is equipped with tapered roller bearings on each shaft. The forces on each bearing must be calculated based on the assumed thermal rating of 105 hp. Using simple beam equations, the forces on each bearing can be solved. Please note that the forces listed are the equivalent radial loads and include the combined effects of both radial and thrust loads on each shaft.

Table C.l gives both the operating and geometric characteristics for each bearing that is necessary for calculating bearing power losses.

The bearing power losses must be calculated for each bearing individually and then summed for the total bearing power loss.

The load dependent power losses for each bearing are calculated using equations 10 and 11. The load dependent losses for bearing #l follows:

Calculate rolling torque, Tb, using equation 11. The coefficient of friction, fb, comes from table 2.

T -fbwJ + 4) b- 4

T _ (0.0018) (1182) (5.875 + 3.000) b- 4

Tb = 4.72 lb in

Next calculate the power loss, PB, for bearing #l using equation 10.

Tbnb pB = 63025

pB _ (4.72) (1780) 63025

PB = 0.133 hp

ANSIJAGMA 601 O-F97

Table C.l - Bearing operating conditions and iecimetry

nput shaft Shaft speed, rpm (?Zb) Bearing #l

Outside diameter (4) Bore diameter (4) Bearing “K-factor” (KS) Bearing length through bore (8) Mean roller diameter (&) Equivalent radial load (IV)

Bearing #2 Outside diameter (4) Bore diameter (4) Bearing “K-factor” (KS) Bearing length through bore (8) Mean roller diameter (DR) Equivalent radial load (IV)

Intermediate shaft Shaft speed, rpm (Q) Bearing #3

Outside diameter (4) Bore diameter (4) Bearing “K-factor” (KS) Bearing length through bore (6) Mean roller diameter (&) Equivalent radial load (w)

Bearing #4 Outside diameter (4) Bore diameter (4) Bearing “K-factor” (KS) Bearing length through bore (5) Mean roller diameter (&) Equivalent radial load (IV’)

Output shaft Shaft speed, rpm (a) Bearing #5

Outside diameter (4) Bore diameter (4) Bearing “K-factor” (KS) Bearing length through bore (8) Mean roller diameter (&) Equivalent radial load (IV)

Bearing #6 Outside diameter (4) Bore diameter (4) Bearing “K-factor” (IQ) Bearing length through bore (8) Mean roller diameter (&) Equivalent radial load (IV)

1780

5.875 in 3.000 in 1.61 2.135 in 3.66 in 1182 lb

4.063 in 1.938 in 1.97 1.751 in D.48 in 1008 lb

1483

5.000 in 2.125 in 1.96 2.063 in D.66 in 2367 lb

4.813 in 2.625 in 1.73 1.510 in 0.58 in 1477 lb

349

8.375 in 4.500 in 1.79 2.625 in 0.92 in 2433 lb

6.375 in 3.000 in 1.46 2.169 in 0.66 in 1618 lb

43

ANSVAGMA 601 O-F97 AMERICAN NATIONAL STANDARD

In a like fashion the torques and power losses can be calculated for bearings #2 through #6.

Torques (Tb) Bearing #2 2.72 lb in Bearing #3 7.59 Ibin Bearing #4 4.94 lb in Bearing #5 14.10 lb in Bearing #6 6.82 lb in

Power losses (PB) Bearing #2 0.077 hp Bearing #3 0.179 hp Bearing #4 0.116 hp Bearing #5 0.078 hp Bearing #6 0.038 hp

The total load dependent power loss is the sum of the six individual bearing load dependent power losses.

; (PB), = 0.133 + 0.077 + 0.179 + 0.116 II= 1

+ 0.078 + 0.038 = 0.621 hp

The non-load dependent losses for bearings are calculated only for tapered roller bearings. For all other styles of bearings, the non-load dependent losses are included in the load dependent loss calculation.

The non-load dependent losses for each bearing are calculated using equations 23, 24 and 25 or 26. The non-load dependent losses for bearing #1 are calculated as follows:

First solve for the arrangement constant, A, using equation 23. Since it was stated that the reducer is lubricated using IS0 220 m ineral oil and since the maximum oil sump temperature is 200” F, then the absolute oil viscosity at maximum oil sump temperature will be 16 centipoises.

A-22440 -- CP

..,

A = y = 1403

Next solve for the cup angle, a~, for the bearing. Since the cup angle was expressed in this example as a "K-factor", KS, use equation 26.

aB . .

aB = 13.58"

426)

Now solve for the non-load dependent (windage and churning) losses, Pm, for bearing #l using equation 24.

P,= d2, n2 B COS3 ag D, 126 000 x (0.78)A

where: do + d.

dm = --+

. ..(24)

dm = 5.875 ; 3.000 = 4.4375in

P - (4~~)2(1780)2(2.135)(cos13.580)3(0.66)

(126 OOO)(n)(O.78)(1403)

PwB = 0.186 hp In a like fashion the non-load dependent losses, Pm, can be calculated for bearings #2 through #6.

I Power losses (PWB) 1 Bearing #2 0.052 hp Bearing #3 0.083 hp Bearing #4 0.057 hp Bearing #5 0.026 hp Bearing #6 0.008 hp

The total non-load dependent power loss is the sum of the six individual bearing non-load dependent power losses.

; (Pm) = 0.186 + 0.052 + 0.083 + 0.0 ?Z= 1 ; 0.026 + 0.008 = 0.412 hp

C.5.1.2 Gearing power losses, PM and PWG The mesh power loss, PM, must be calculated based on the assumed thermal rating of 105 hp since the coefficient of friction is dependent upon the contact load factor, K. Table C.2 which follows provides the gear geometry characteristics which must be deter- m ined prior to calculating the gearing power losses, PM and %o. The gearing power losses must be calculated for each mesh individually and then summed for the total gear power loss. The load dependent power losses for each mesh are calculated using equation 12. Equations 13 through 17 are also required to calculate some of the factors used in equation 12. The load dependent losses for the first stage follow: The first step is to calculate the mesh mechanical advantage, M , using equation 15. This calculation requires the solution of the sliding ratio at the start of approach, H,, equation 16, and the sliding ratio at the end of recess, Ht, equation 17.


Table C.2 - Gear geometry characteristics Sliding ratio at start of approach:

=irst stage Pinion teeth Np =30 Gear teeth &=36 Gear ratio mG= 1.20 Center distance = 7.250 in Transverse operating &, =24.62721" pressure angle Pinion outside radius r0 = 3.509 in Gear outside radius & =4.155 in Pinion operating pitch radius r, = 3.295 in Gear operating pitch radius & = 3.955 in Pinion speed, rpm np =1780 Pinion torque Tp =3718Ibin Effective face Fe = 1.750 in Pitch fine velocity V = 3071 fpm Operating helix angle wy = 19.77250” Pinion operating pitch d,,, = 6.591 in diameter Gear operating pitch D, = 7.909 in diameter Pinion total face Ft = 2.125 in Gear total face Ft = 1.750 in Normal diametral pitch ptl =5

second stage Pinion teeth Np =12 Gear teeth N~=51 Gear ratio w = 4.25 Center distance = 7.250 in Transverse operating +,+, = 23.74476" pressure angle Pinion outside radius r, = 1.670 in Gear outside radius & = 6.040 in Pinion operating pitch radius r,.,, = 1.381 in Gear operating pitch radius & = 5.869 in Pinion speed, rpm np =1483 Pinion torque Tp = 4372 lb in Effective face Fe = 5.000 in Pitch line velocity V = 1073 fpm Operating helix angle wy = 8.30276" Pinion operating pitch d, =2.762 in diameter Gear operating pitch D, = 11.738 in diameter Pinion total face 4 = 6.040 in Gear total face Ft = 5.000 in Normal diametral pitch PFl = 4.5

Hs=(mG+l)[(g - cos2&~- h&]

0.5 II, = (1.20 + 1) cm2 24.62721’

- sin24.62721”] = 0.242

Sliding ratio at end of recess:

. ..(16)

- sin24.62721"] = 0.253

Mesh mechanical advantage:

M = 2-hv(H, + 4) fe+H:

M = 2cos24.62721"(0.242 + 0.253) 0.2422 + 0.2532

. ..(15)

M = 7.342 The second step is to calculate the mesh coefficient of friction, fm, using equation 13. This calculation also requires the solution of the contact load factor, K, using equation 14.

Contact load factor:

K _ TP(NP +NG)

2 F (6) NG

K= 3718(30 + 36)

2(1.750)(3.295)'(36) K = 179.3 lb/in2

. ..(14)

Mesh coefficient of friction:

f = I@-35 m c, vo.23

ftn = 179.30-35 (80.1)(3071)“-23

. ..(13)

fm = 0.012

where:

Cl (from table 3) = 80.1

45


Now calculate the load dependent mesh power loss:

P,= fm Tp np ~09 V W 63025 M

p M

= (0.012)(3718)(1780) cos2 19.72250" 63 025(7.342)

P M = 0.152 hp . ..(12)

In a like fashion, using the same equations, the load dependent losses for the second stage of gearing can be calculated:

H, = 0.356 Ht = 0.479 M = 4.292

i = 283 lb/in2 = 0.018

PM = 0.425 hp

The total load dependent power loss is the sum of the two individual gear stage load dependent power losses:

;: PM), = 0.152 + 0.425 n=l

i PM), = 0.577 hp n=l

The non-load dependent losses (windage and churning), PWG, for the gearing are calculated for each gear and pinion individually using equations 19 and 20 and then summed for the total non-load dependent gear loss. Note that the arrangement constant, A, equation 23, is the same value as was previously calculated for the bearings.

Non-load dependent loss for the first stage pinion:

P,= d$, n2 Ft cm3 I),.,,

126 000 P,A

P wP= (6.591)2(1780)2(2.125) cos3 19.77250”

(126 000)(5)(1403)

PUT = 0.276 hp . ..(20)

In a like fashionthe non-load dependent losses can be calculated for the other gears and pinions in the gear train:

First stage gear 0.227 hp Second stage pinion 0.124 hp Second stage gear 0.102 hp

The total non-load dependent power loss for the gears and pinions:

; (PwG), = 0.276 + 0.227 + 0.124 + 0.102 n=l

4 c (P&n = 0.729 hp

n=l --

C.5.1.3 Oil seal power loss, Ps

The gear reducer in this example has a single extended input shaft with a single 3.000 inch diameter BUNA-N oil seal and a single extended output shaft with a single 4.500 inch BUNA-N oil seal. The input shaft rotates at 1780 rpm and the output shaft rotates at 349 rpm.

Equation 18 is used to calculate the oil seal power losses.

Input shaft oil seal power loss:

Ts n " = 63025 Ts (from figure 1) = (0.536)(3) T, = 1.608 lb in p

S

= (1.608)(1780) 63025

ps = 0.045 hp

. ..(18)

In a like fashion the oil seal power loss for the output shaft can be calculated to be:

P, = 0.013 hp The total oil seal power loss for the gear reducer is:

Ii (Ps)n = 0.045 + 0.013 n=l

2 2 (Ps)n = 0.058 hp

n=l

C.5.1.4 Total heat generated, PV

The total heat generated, Pv, in the gear drive is the total of the load dependent losses, PL, and the non-load dependent losses, PN. Use equations 2,3, 8 and 9 to total the heat generated, fi, as follows:

6 2 PL = c w, + 2 PM),

n=l n=l

PL = 0.621 + 0.577

PL = 1.198 hp . ..(8)

46


PN = ps + CPWC + CPWB + cpp

PN = 0.058 + 0.729 + 0.412 + 0 PN = 1.199 hp PV=PL +PN Pv = 1.198 + 1.199 Pv = 2.397 hp

. ..(9)

. ..(3)

C.5.2 Heat dissipation, PQ

The gear reducer has a cast iron gear case with a surface available for heat transfer of 21.8 ft2 . The heat dissipation is calculated using equation 30.

PQ =A, kAT 4 = 21.8 ft2

. ..(30)

k (from table 4) = 0.0011 hp/(ft2 OF) AT = 100°F PQ = (21.8)(0.0011)(100) PQ = 2.398 hp

Since the gear drive is in thermal equilibrium, PQ = Pv, the assumed thermal rating of 105 hp becomes the calculated thermal capacity, PT. if through this calculation method, PV # PQ then another assumption must be made for the thermal rating. This new assumption must be used to recalculate the load dependent losses,

PL = CPB + CPM’ and additional assumptions

made for the thermal rating until such time as PV=PQ.

C.5.3 Efficiency, q

The efficiency is calculated using equation 6.

rl = 100 ‘L +‘N - PA

x 100

rl = 1oo- 1.198 + 1.199 x 1oo 105 . ..(6)

?j = 97.717% The thermal capacity of the gear drive is calculated using equation 7.

P, = pQ

P, = 2.398 1 97.717 .,

100 -(7)

P, = 105 hp

Note that the calculated thermal power rating, PT, is the same value as the assumed thermal power rating under conditions of thermal equilibrium.

47

ANSI/AGMA 6010-F97


48

AMERICAN NATIONAL STANDARD ANSIJAGMA 601 O-F97

Annex D (informative)

Test and inspection procedures

phe foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANWAGMA 6010-F97, Standard for Spur, Helical, Herringbone and Bevel Enclosed Drives.]

D.l Purpose

This annex covers the testing and inspection procedures for assembled gear drives. individual component inspection and process control are beyond the scope of this standard.

When testing of the gear drive is required, the drive should be properly mounted for running the test in the intended operating position to ensure that all facets of the assembly are correct. Under normal test conditions the gear drive is connected by coupling or belt drive to an electric motor that is available for the purpose at the manufacturer’s test facility. The following applies to only those gear drives which are lubricated in accordance with manufacturer’s recommendations and tested in a system of connected rotating parts. During testing, the system should be free from critical speeds, torsional vibrations and overloads as tested at the gear drive manufacturer’s facility.

D.2 Inspection of the assembled gear drive

The correct mating of a gear set depends not only on the accuracy of the gear teeth, but also on the position and the alignment of the gear axes relative to each other. The components, having been fully approved prior to assembly, are assembled, and proper tooth contact, backlash and bearing settings are verified.

D.2.1 Tooth contact inspection

Checking the tooth contact pattern (tooth bearing area) is frequently an important test of the gear drive and is of special value when gears have been mounted in a housing, because the test will indicate if the helix and pressure angles and the resultant base pitch of the mating gears meet the specified requirements and achieve optimal gear performance. The pinion profiles are generally coated with a marking compound and then rotated in mesh with the mating gear, and the resulting tooth pattern can be docu- mented. See AGMA 390.03a, AGM4 Handbook - Gear Classification, Materials and Measuring Meth- ods for Bevel, Hypoid, Fine Pitch Wormgearing and

Racks only as Unassembled Gears, 1980, Part II I, Section 9, “Tooth Contact Pattern” and ANSVAGMA 2000-A88, Gear Classification and Inspection Handbook - Tolerances and Measuring Methods for Unassembled Spur and Helical Gears (Including Metric Equivalents), 1988, Appendix D, “Contact Pattern Check”.

The percentage of tooth contact will vary depending upon the loading of the gears, but the pattern obtained even under a no load condition will provide the manufacturer with important information.

D.2.2 Backlash

Backlash in gears is the clearance or play between mating tooth surfaces. The backlash will be a function of the tolerances on tooth thickness, runout, lead, profile, center distance, and by the temperature differences between the housing and the gears. Functional backlash is the backlash at the tightest point of mesh on the pitch circle in a direction normal to the tooth surfaces when the gears are mounted in their assembled positions.

Backlash is typically measured with feeler gauges or dial indicators normal to the gear tooth for a given mesh.

Circumferential backlash of the assembled unit with gears other than spur gears should take into account the axial float of the shafts involved.

D.2.3 Rolling element bearings

When rolling element bearings are used, the manufacturer, based on his experience, the application, and the recommendations of his bearing supplier, will determine the type of bearings and their settings. Assembly procedures normally require a tolerance to be established for the desired setting. An incorrectly set bearing can be a source of damage for the gear drive. Bearing end play may be set one shaft at a time and finally checked when both end cover plates are bolted in place with the required shims. End play should be checked to ensure compliance with the specification. Full end play is typically measured with the shaft moved all the way

49

ANSI/AGMA WlO-F97 AMERICAN NATIONAL STANDARD

in one direction and then moved fully in the other direction. Total movement is the end play.

0.3 Testing procedure

For the purpose of a running test the following conditions would apply.

D.3.1 Speed

A gear drive intended for service at a single speed shall be tested at that speed unless otherwise agreed upon between gear manufacturer and purchaser. The test speeds of a gear unit intended for service over a range of operating speeds shall span the range of operating speeds, unless otherwise negotiated between the manufacturer and the purchaser. The direction of rotation during the test shall be the same as that intended in service, if known.

D.3.2 Loading

Gear drives may be operated with or without load at the gear manufacturer’s discretion unless specific test loads are agreed upon and included as a part of the purchase contract. In individual cases, especially where unusually high speeds or power are involved, alternate operating conditions may be negotiated.

CAUTION: It is recommended that gear drives not be tested with loads in excess of unit rating, since such practice will reduce the design life of the unit.

D.3.3 Test requirements

The duration of the running test will be decided by the drive manufacturer unless a specific time has been contractually agreed upon between manufacturer and purchaser.

Features such as oil tightness, noise level, temperature rise, axial and radial play of input and output shafts, contact pattern of the gear meshes, and lubrication system may be checked and recorded at this time.

CAUTION: It is recommended that gear drives not be tested with loads in excess of gear unit rating, since such practice will reduce the design life of the unit.

D.3.4 Lubrication system performance

The lube system must be checked for adequacy at certified speed or at both ends of speed range if the speed is variable:

- On splash systems, the oil level must be high enough to lubricate all components. It must not be unnecessarily high because sound and heat will be generated;

- On pressure lube systems, oil lines, troughs, gauges, pumps, filters, etc., must be checked for performance and any leakage. Flow, pressure, and temperature are to be recorded at regular intervals.

D.3.5 General

- Any deviations from any applicable specifications on the certified print will be noted on the test report;

- All deficiencies such as oil leaks, excessive sound level, vibration, abnormal temperature rise, and insufficient tooth contact must be corrected before the gear drive is shipped;

- The ratio should be verified along with the assembly, shaft extension details, and direction of rotation.

50


Annex E (informative)

Owner responsibilities

rheforeword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSVAGMA 6010-F97, Standard for Spur, Helical, Herringbone and Bezel Enclosed Drives.]

This annex lists which applicable items must be considered and properly provided for by the owner.

Many different types of units for widely varying applications are covered by this standard. This section, which is arranged as a checklist, is intended to act as a guide. Specific items should be applied as appropriate for the particular unit for the specific application.

E.l Specifications

Owner has the responsibility to specify to the manufacturer such items as the required loads and the operating environment.

E.2 Storage and handling

- Proper storage of unit until installed;

- Proper preservation of the unit until it is placed into service;

- Proper handling of the unit:

- safety of personnel comes first;

- lift only at adequate lifting points;

- protect the mounting surface from damage.

E.3 Installation

- Proper installation of unit on an adequate foundation:

- adequately supported;

- securely bolted into place;

- properly leveled so as not to distort the gear case.

- Properly install couplings suitable for the application and connected equipment;

- Ensure accurate alignment with other equipment;

- Furnish and install adequate machinery guards as needed to protect operating personnel and as required by the applicable standards of the Occupational Safety and Health Administration (OSHA), and by other applicable safety regulations;

- Ensure that driving equipment is running in the correct direction before coupling to gear drive designed to operate in a specific direction.

E.4 Start-up

- Ensure that switches, alarms, heaters, coolers and other safety and protection devices are installed and operational for their intended purposes;

- On a unit equipped with a separately driven lubrication pump, run the pump and check out the lubrication system prior to starting the unit;

- Fill the unit or sump to proper level with correct lubricant before starting drive. Refill as necessary immediately after starting the unit;

- Ensure that all grease points have received the proper amount of grease.

E.5 Operation and maintenance

- Operate the equipment as it was intended to be operated:

- do not overload;

- run at correct speed.

- Maintain lubricant in good condition and at proper level;

- Dispose of used lubricant in accordance with applicable laws and regulations;

- Apply proper amount of grease to specified locations at prescribed intervals;

- Perform periodic maintenance of the gear drive as recommended by the manufacturer.

51



52

AMERICAN NATIONAL STANDARD ANSIjAGMA 601 O-F97

Annex F (informative)

Gear tooth mesh losses for bevel gears

rhe foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSVAGMA 6010-F97, Standard for Spur, Helical, Herringbone and Bevel Enclosed Drives.]

The following procedure for determining the heat generated by bevel gears is a new procedure; therefore, it is recommended that testing be included to verify results.

The following equations can be used to estimate the gear tooth mesh losses, Pm, for bevel gears. See figures F.l and F.2 for either taper or untform depth tooth.

Figure F.1 - Uniform depth tooth

. . . (F.1)

where

fm is coefficient of friction (see equation 13);

TP is pinion torque, lb in;

9 is pinion speed, rpm;

q is mean spiral angle, degrees;

M is mesh mechanical advantage.

M = 2 cm cPr (& + 4)

e+e

Figure F.2 - Taper depth tooth

where

4 is transverse pressure angle.

&= tan-* g$ ( )

4) is normal pressure angle.

For bevel gearing, pitch line velocity, v, used in equation 13, is calculated at large end of tooth.

The K-factor is given by the equation:

K = TP (NP + 47) 2F rj$, NC

. ..(F.2)

where

Np is the number of pinion teeth;

No is the number of gear teeth;

F is face width in contact with mating element, in;

rm is mean reference radius, pinion, in.

53

ANSI/AGYA 6010-F97 AMERICAN NATIONAL STANDARD

The values for Hs and Hr are:

L

...(F.3)J

where

. ..(F.4)

w is equivalent gear ratio; re,,,p is equivalent mean reference radius, pinion,

in; r,,G is equivalent mean reference radius, gear,

in; reo,,p is equivalent tip radius at m id-face width,

pinion, in; reO,,,Gis equivalent tip radius at m id-face width,

gear, in. The equation for equivalent mean reference radius is given by:

Am r rem = Ao CQS Y

. ..(FS)

where

r is pitch radius, in;

AtI is mean cone distance, in; & is outer cone distance, in;

Y is reference cone angle. The equivalent gear ratio can be calculated as:

‘emG meG = ‘snp . . . F.6)

r&m = Gm + &I where

. ..(F.7)

am is mean addendum, at m id-face, in. If the addendum at outer end and the face angle are known, the addendum at m id-face can be calculated as:

a,=a - F tduo - Y) 2

where

F is face width, in; a is addendum at outer end, in;

Yo is face angle; for uniform depth teeth y. = y;

Y is reference cone angle. Equations F.5, F.7 and F.8 are to be calculated for both pinion and gear member using respective pitch radius, r; pitch angle, y; mean addendum, h; and face angle, yo.

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AMERICAN NATIONAL STANDARD ANSVAGMA 6010-F97

Bibliography

rhe foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSIJAGMA 6010-F97, Standard for Spur, Helical, Herringbone and Bevel Enclosed Drives.]

1. Townsend, Dennis P., Dudley’s Gear Hand- book, Second edition, McGraw-Hill, New York, 1992.

5. ANSIIAGMA 6032-A88, Standard for Marine Gear Units: Rating.

2. The Timken Company: Bearing Torque, Heat Generation and Operating Temperature.

6. Winter, H. and Michaelis, K., Scoring Load Capacity of Gears Lubricated with EP-Oils, AGMA Technical Paper P219.17.

3. Palmgren, Dr. Eng.: Ball Roller Bearing Engi- neering, Third Edition.

7. Dudley, Dane E.: Solar Test Report, by M. Dunn (also ANSI/AGMA 6032-A94, equation 9.19).

4. Eschmann, Hasbargen and Weigand, Ball Roller Bearings, Theory, Design and Application, Second edition, John Wiley and Sons, Ltd., Chi- Chester, 1985.

8. Rubber Manufacturers Association: Techni- cal Bulletin OS-l 5.

9. McAdams, Will iam H., “Heat Transmission”, Third Edition, Chapter 9.

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ANSI/AGMA 601 O-F97

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Ansi Agma6010 f97

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