Y+
l
& ..= -XB
Reproduced By G L O B A L E N G t N E E R l N GD O C U M E N T S c With T h e P e r m i s s i o n O f A G M A U n d e r Royalty A g r e e m e n t
AGMA911-A94
i
c
A M E R IC A N G E A R M A N U F A C T U R E R S A S S O C IA T IO N
D e sig n G u idelines fo r A e r o s p a c e G e a rin g
c.
AGMA
INFO R M A T IO N
SHEET
(This In fo r m a tio n S h e e tis N O T a n A G M A S ta n d a r d )
AGMA 911-A94
AGMA 91%A94, Design Guidelines for Aerospace Gearing CAUTION NOTICE: AGMA standards 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 Association on the subject matter. [Tables or other self-supporting sections may be quoted or extracted in their entirety. Credit lines should read: Extracted from AGMA 911-A94, information Sheet - Design Guidelines for Aerospace Gearing, with the permission of the publisher, the American Gear Manufacturers Association, 1500 King Street, Suite 201, Alexandria, Virginia 223141.
ABSTRACT: This Information Sheet covers current gear design practices as they are applied to air vehicles and spacecraft. The material included goes beyond the design of gear meshes and presents the broad spectrum of factors which combine to produce a working gear system, whether it be a power transmission or special purpose mechanism. Although a variety of gear types, such as wormgears, face gears and various proprietary tooth forms are used in aerospace applications, this document covers only spur, helical, and bevel gears.
Copyright 0 1994 by American Gear Manufacturers Association Published by American Gear Manufacturers Association 1500 King Street, Suite 201, Alexandria, Virginia, 22314
ISBN:
ii
l-55589-8294
AGMA 911-A94
Contents Foreword . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ..*...............
Page vi
1
Scope .......................................................................
1
1.1 1.2
Application ................................................................... References ...................................................................
1 1
2
Application ...................................................................
1
3 3.1 3.2
Definitions and symbols ........................................................ Definitions ................................................................... Symbols .....................................................................
2 2 2
4 4.1 4.2 4.3 4.4
Designapproach .............................................................. Design requirements and goals ... ;. ............................................ Identify design criieria ......................................................... . ....................... Preliminary design .................................... ................................................................ Detail design
5 5 6 8 12
5 5.1 5.2 5.3 5.4 5.5 5.6 5.7
Lubrication .................................................................. Cooling vs. lubrication requirements ............................................ Lubricants...................................................................l Distribution systems .......................................................... Lubrication system design considerations ....................................... Filtration .................................................................... Oiipumps ................................................................... Lube system condition monitoring ..............................................
15 15 5 18 19 21 21 23
6 6.1 6.2 6.3 6.4 6.5 6.6 6.7 6.8
Environmental issues ......................................................... Ambient temperature effects ................................................... Ambient pressure effects ...................................................... Attitude effects .............................................................. Contaminant effects (water, corrosives, dirt, dust, and sand) ....................... Vibration/Shock effects ....................................................... Fire resistance requirements .................................................. Electromagnetic effects ....................................................... :. ....... Nuclear, biological, and chemical (NBC) effects .........................
24 24 25 25 26 26 29 29 29
7 7.1 7.2 7.3 7.4 7.5
Vibration and noise ........................................................... Causes of gear vibration ...................................................... Consequences of vibration .................................................... Design ...................................................................... Analyzing vibration problems .................................................. VibratiorYNoise reduction techniques ...........................................
30 30 31 32 35 37
8 8.1 8.2 8.3
Load Capacity ............................................................... Introduction ............................................................... Spur, helical, and bevel gear tooth breakage and surface durability ................. Spur, helical, and bevel gear scuffing (scoring) -flash temperature index ............
9 9.1 9.2
Gear materials and heat treatment ............................................. Class and grade definitions .................................................... Mechanical properties ........................................................
39 ..3 9 41 45 47 47 47
... III
AGMA 911-A94
Contents, continued 9.3 9.4 9.5 9.6 9.7 9.8 9.9
9.10 9.11 9.12 9.13
Cleanliness ............................................................... ................................... Heat treatment ............................ Microstructure ............................................................... Hardenability ................................................................ Dimensional stability .......................................................... Pm-machining stock removal .................................................. Ferrousgearing .............................................................. Non-ferrous gearing .......................................................... Material grades and heat treatment ............................................. Gear surface hardening ....................................................... Gear through hardening .......................................................
..4 8 48 48 48 48 48 48 49 49 49 53
10.1 10.2 10.3 10.4
Surface treatment ............................................................ Introduction ................................................................. Shot peening ................................................................ Surfacecoatings.............................................................6 Ion implantation of gears ......................................................
54 54 55 0 61
11 11.1 11.2 11.3 11.4
Manufacturing considerations .................................................. Introduction ................................................................. Spur and helical gears ........................................................ Bevel gears ................................................................. Stress relief treatment ........................................................
63 63 63 64 67
12 12.1 12.2 12.3
Gear inspection .............................................................. General ..................................................................... Spur and helical involute gears ................................................ Bevelgears .................................................................
68 68 68 69
13 13.1 13.2 13.3
Rocket & space gearing .... .................................................. Introduction ................................................................. Lubrication .................................................................. Gear materials for space application ............................................
70 70 71 73
1
Symbols used in equations .....................................................
10
Aerospace lu&icant viscosities ................................................ Aerospace lubricant densities .................................................. Aerospace lubricant pressure/viscosity coefficients ............................... Aerospace lubricant specific heat values ........................................ Aerospacegreases ........................................................... Aircraft dry lubricants ......................................................... Advantages & disadvantages of a common engine & transmission lubrication system .
2 16 16 16 17 17 17 19
9 10 11
26 Particle size distribution by weight .............................................. Suggested functional test levels for propeller aircraft and turbine engine equipment ... 27 28 Suggested functional test peak levels for equipment installed on helicopters .........
12 13 14
Potential influence of design features on noise and vibration ....................... Vibration testing ............................................................. Sound and vibration reduction techniques .......................................
iv
35 37 39
AGMA Qll-AQ4
Tables, continued 15
Typical aerospace gear materials ..............................................
49
16
Surface coatings used in aerospacegear units .....................................
62
17 18 19
Candidate solid-film lubricants for space application .............................. Candidate fluid lubricants for space application .................................. Working fluid lubricants .......................................................
73 74 74
Figures Retative life as function of lambda ............................................... The general parallel-axis epicyclic gear train ..................................... Goodman diagram for combined load ...........................................
7 9 11
Typical aerospace lubrication system schematic ................................. Spur gear pump ............................................................. Vane pump .................................................................. Gerotor Pump ...............................................................
22 23 23 23
8 9 10 11
Typical gearbox attitude limits .................................................. Suggested vibration spectra for propeller aircraft and turbine engine equipment ...... Suggested vibration spectrum for equipment installed on helicopters ................ Terminal-peak sawtooth shock pulse configuration and its tolerance limits ...........
25 27 27 28
12 13
Sound and vibration paths ..................................................... Typical damping ring .........................................................
30 38
14 15 16
Diierent methods for determining tooth root stress ............................... Directions of crack propagation in gear teeth .................................... Reliabilii versus number of standard deviations .................................
40 40 46
17 18
Schematic of material ground from a gear tooth .................................. Schematic of material ground from a distorted gear tooth ..........................
53 53
19 20 21 22 23 24 25 26 27 28
Fatigue strength in ground AISI 4349 (50 HRC) .................................. Example of residual stress profile created by shot peening ........................ Peening 1045 steel at 48 HRC with 330 shot .................................... Peening 1045 steel at 62 HRC with 330 shot .................................... Stress profile of carburized gear tooth root, ground and then shot peened ........... Increase in fatigue resistance of spiral bevel gear ................................ Fatigue tests on rear axle shafts ............................................... Fatigue tests on notched shafts ................................................ Fatigue life comparison ....................................................... Correlation of Almen intensities as indicated by arc heights of peened strips .........
54 55 56 56 57 57 57 57 58 59
29
Heat treat coupon ............................................................
68
1 2 3
Annexes A Spur gear geometry factor including internal meshes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 75 Gearbox test and mission requirements . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 89 B C References and bibliography . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 97
V
AGMA 911-A94
[The foreword, footnotes, and annexes are provided for informational purposes only and should not be construed as a part of AGMA 911-A94, information Sheet - Design Guidelines for Aerospace Gearing,] This Information Sheet supersedes AGMA Standard 411.02, Design Procedure forAirm&Engine and Power Take-OlT Spur and Helical Gears. Its purpose is to provide guidance to the practicing aerospace gear engineer in the design, manufacture, inspection, and assembly of aerospace gearing. In addition, it addresses the lubrication, environmental, and application conditions which impact the gearbox as a working system of components. Material in the Information Sheet is supplemental to current AGMA Standards, but does not constitute a Standard itself. By definition, Standards reflect established industry practice. In contrast, some of the practices discussed here have not seen enough usage to be considered standard, but they do provide insight to design techniques used in stat-f-the-art aerospace equipment. It is expected that the user of this Information Sheet will have some general experience in gear and machine design, and some knowledge of current shop and inspection practices. Suggestions for the improvement of this information sheet will be welcome. They should be sent to the American Gear Manufacturers Association, 1500 King Street, Suite 201, Alexandria, Virginia, 22314.
vi
AGMA 911-A94
PERSONNEL of the AGMA Committee for Aerospace Gearing Chairman: A. Meyer . . . . . . . . . . . . . . . Textron - Lycoming Vice Chairman: K. Buyukataman . . . . Pratt & Whitney
ACTIVE MEMBERS J. Abrahamian . . . Pratt & Whitney N. Anderson . . . . . GM Technical Center I. Armitage . . . . . . Spar Aerospace E. J. Bodensieck . Bodensieck Engineering M. Brogiie . . . . . . . Dudley Technical Group R.C. Bryant . . . . . . General Electric Ft. Burdick . . . . . . . Aero Gear J. Daly . . . . . . . . . . Metal Improvement Co. R. Dayton . . . . . . . Wright Patterson A. F. 9.
R. Drago . . . . . . . . Boeing Helicopters B. Dreher . . . . . . . Kaiser Aerospace R. C. Ferguson . . Taiga Group W. D. Glasow . . . . Sikorsky T. Heiliger . . . . . . . Sikorsky M. Howes . . . . . . . IIT Research J. G. Kish . . . . . . . Sikorsky E. A. Lake . . . . . . Wright Patterson A.F.B. W. Michaels . . . . . Sundstrand
ASSOCIATE MEMBERS G. Belling . . . . . . . American Pfauter J. D. Black . . . . . . General Motors E. R. Braun . . . . . Eaton C. E. Breneman . Advance Gear A. T. Brunet . . . . . Allied Signal Aerospace J. Cadisch . . . . . . Reishauer H. S. Cheng . . . . . Academic Member L. Cloutier . . . . . . Academic Member 9. Cluff . . . . . . . . . American Pfauter F. W. Cumbow . . . M&M Precision R. J. Cunningham Boeing P. A. Deckowitz . . ITWlllitron J. W. Dern . . . . . . SPECO Corporation K. R. Dirks . . . . . . Allied-Signal, Garrett Eng. Div. R. DiRusso” . . . . . Kaman D. W. Dudley* . . . Honorary Member R. Durwin . . . . . . . Sikorsky W. C. Emmerling Naval Air Propulsion Center R. L. Errichello . . Academic Member J. A. Ferrett . . . . . National Broach D. J. Fessett . . . . Lucas Western IncJATD H. K. Frint . . . . . . Sikorsky R. Gefron . . . . . . . Superior Gear N. L. Grace . . . . . Gleason Works M. J. Gustafson . Kaman D. R. Houser . . . . Academic Member C. lsabelle . . . . . . Sikorsky D. E. Kosal . . . . . National Broach C. Layer . . . . . . . . Mmg A. J. Lemanski Academic Member A. A. Lewis . . . : : Pratt & Whitney, Canada M. Lonergan . . . . National Broach P. Mangione . . . . Naval Air Warfare Center W. Mark . . . . . . . . Academic Member
W. Marquadt * . . . Norwood Precision/Textron D. Merritt . . . . . . . Lion Precision Gear R. Miller . . . . . . . . Pratt &Whitney J. Mogul* . . . . . . . Metal Improvement Co. J. O’Donnell . . . . . Naval Air Warfare Center A. E. Phillips . . . . Emerson Power Transmission T. L. Porter . . . . . . ITW/Spiroid A. K. Rakhit* . . . . Solar Turbines J. R. Reed . . . . . . Klingelnberg Soehne T. Riley . . . . . . . . . NWL Control System E. Ropac . . . . . . . Bachan Aerospace S.S. Sachdev . . . Spar Aerospace B. Schneider . . . . NASA, Johnson Space Center D. J. Schreiner . . General Motors A. Seireg . . . . . . . Academic Member S. V. Shebelski . . Lion Precision Gear E. E. Shipley . . . . Mechanical Technology G. Skirtich . . . . . . Lion Precision Gear L. J. Smith . . . . . . Invincible Gear N. Sonti . . . . . . . . Academic Member D. A. Sylvester . . Power-Tech K. Tower . . . . . . . . Metal Improvement Company D. P. Townsend* . NASA, Lewis F. Uherek . . . . . . . Flender M. Valori . . . . . . . . Naval Air Propulsion Center L. Vesey . . . . . . . . iTW/Spiroid D. A. Wagner . . . . General MotorsIAGT H. Wagner . . . . . . Advance Gear R. D. Wagner . . . National Broach 9. R. Walter . . . . . Liebherr Machine R. F. Wasilewski . Arrow Gear S. R. Winters . . . . General Motors T. J. Witheford . . . Teledyne Vasco G. I. Wyss . . . . . . Reishauer
* Contributed technical material to the document.
vii
This page is intentionally blank.
. .. VIII
AGMA 911-A94
AGMA 91%A94
Design Guidelines for Aerospace Gearing 1 Scope This Information Sheet covers current gear design practices as they are applied to air vehicles and spacecraft. The material included goes beyond the design of gear meshes per se, and presents, for the consideration of the designer, the broad spectrum of factors which combine to produce a working gear system, whether it be a power transmission or special purpose mechanism. Although avariety of gear types, such as wormgears, face gears and various proprietary tooth forms are used in aerospaceapplications, this document covers only conventional spur, helical, and bevel gears.
1 .l Application The working environment of the aerospace gear has become so diverse that a single set of guidelines will no longer suffice. The operating conditions imposed on a high speed, high powered, transmission or actuator are quite different than those experienced by the spacecraft mechanism which must function in a hard vacuum for long periods of time without maintenance. This Information Sheet addresses these differences and provides guidance to the designer for these demanding applications.
1.2 References The following standards contain provisions which, through reference in thii text, constitute provisions of this American Gear Manufacturers Information Sheet. At the time of publication, the editions indicated were valid. All standards are subject to revision, and parties to agreements based on this American Gear Manufacturers lnfomration Sheet are encouraged to investigate the possibility of applying the most recent editions of the standards indicated below. AGMA 230.01 - 1974, Surface Temper Inspection Process.
AGMA 390.03a, - 1980, Gear Handbook - Gear Classification, Materials and Measuring Methods for Bevel, Hypoid, Fine Pitch Wormgearing and Racks Only as Unassembled Gears. ANSVAGMA 110.04- 1989, Nomenclature of Gear Tooth Failure Modes. ANSVAGMA 2000-A88, Gear Classification and inspection Handbook - Tolerances and Measuring Methods for Unassembled Spur and Helical Gears (Including Metric Equivalents). ANSVAGMA 2001-988, Fundamental Rating Factors and Calculation Methods for Involute Spur and Helica/ Gear Teeth. ANSVAGMA 2003-A86, Rating the Pitting Resistance and Bending Strength of Generated straight Bevel, ZEROLB Bevel, and Spiral Bevel Gear Teeth. ANSVAGMA 20044389, Gear Materials and Heat Treatment Manual ANSVAGMA 6023-A88, Design Manual for Enclosed Epicyclic Gear Drives. ANSVAGMA 6123-A88, Design Manual for Enclosed Epicyclic Metric Module Gear Drives.
2 Application A listing of aerospace geared applications by type of service or function performed is useful in segregating the diverse gearing tasks into mechanism families which experience similar load and environmental spectra. Applications can be identified by general grouping as follows:
- Main propulsion systems; - Propeller gearboxes reduce
engine
speed to propeller speed;
- Fan gearboxes allow the use of optimum turbine and fan speeds for maximum efficiency; - Helicopter transmissions. A system of gearboxes and shafting to drive the helicopter rotors from the engine(s);
AGMA 246.02A - 1983, Practice for Carburked
- Mechanical interconnection between engines allow for independent engine opera-
Aerospace Gearing.
tion on multiingine
systems; 1
AGMA 911-A94
- Accessory drive gearboxes driie accessory devices, such as generators, fuel pumps, hydraulic pumps, oil pressure and scavenge pumps, blowers, alternators, etc; - Auxiliary/secondary power units (APU/ SPU) consist of an auxiliary turbine engine integrated with a gearbox to provide powerfor main engine starting, electrical services, emergency hydraulic power, cabin air conditioning, etc.; - Actuators. A general class of geared devices used to cause a position change of an object. The objects may include aerodynamic control surfaces, winch cables, doors, landing gear, or telerobotic arms. Actuators are distinguished from most aerospace gearing in that they only move on command; - Space systems. A specialized grouping of power (as in rocket turbo-pump drives), and actuatortype devices which have been designed to be compatible with the unique rigors of outerspace environments. These include the high power, short life rocket applications as well as the long life satelliie or space platform systems. 3 Definitions
The terms used, wherever applicable, conform to the following standards: ANSI Y10.3-1968, Letter Symbok for Quantities Used in Mechanics of Solids AGMA 1012-FQO, Gear Nomenclature, Definitions of Terms with Symbols AGMA 904-689, Metric Usage 3.2 Symbols. The symbols used in this information sheet are shown in table 1. NOTE - The symbols and definitions used in this information sheet may differ from other AGMA publications. The user should not assume that familiarsymbols can be used without a careful study of these definitions.
SI (metric) units of measure are shown in parentheses in table 1 and in the text. Where equations require a different format or constant for use with SI units, a second expression is shown after the first, indented, in smaller type, and with “M” included in the equation number. Example S
wt *=K,K,
&
pd
Ks
%
KB
J
w,Ka 1 4 KmKB St=--
and symbols
K~K~ mF J The second expression uses SI units.
3.1 Definitions.
. ..(n) . ..(llM)
Table 1 - Symbols used in equations Symbol
2
Units
Name
Reference First equation paragraph
C
Center distance
in (mm)
9
8.2.2
G
Application factor for pitting resistance
----
12
8.2.2
Cf
Surface condiiion factor for pitting resistance
--a-
12
8.2.2
CH
Hardness ratio factor for pitting resistance
-w--
18
8.2.8
CL
Life factor for pitting resistance
18
8.2.8
Gl
Load distribution factor for pitting resistance
12
8.2.2
cp
Elastic coefficient
12
8.2.2
CR
Reliability factor for pitting resistance
18
8.2.8
G
Size factor for pitting resistance
12
8.2.2
AGMA 91%A94
Table 1 (coMwo” Name
Symbol
Units
First equation
Reference paragraph
CT
Temperature factor for pitting resistance
----
18
8.2.8
G
Dynamic factor for pitting resistance
----
12
8.2.2
ch
Lubricant specific heat
BTlJ/lbm “F
8
5.1.2
&J&l K) dp
Pinion operating pitch diameter
in (mm)
10
8.1.2
E, F
Reduced modulus of elasticity
IWir?(N/mm2)
4
4.2.4
Face width
in (mm)
9
8.1.2
Fe H
Effective or net face width
in (mm)
12
8.2.2
Oil film thickness
in (mm)
1
4.2.4
HG
Heat generated at design point
BTU/min
8
5.1.2
(kJ/min) &ill
Film thickness, minimum
----
4
4.2.4
I
Geometry factor for pitting resistance
----
12
8.2.2
J
Geometry factor for bending strength
----
11
8.2.1
K
Contact load factor for pitting resistance
lb/in* (MPa)
9
8.1.2
Ka
External application factor for bending strength
. - - - -
11
8.2.1
KB
Rim thickness factor
----
11
8.2.1
KL
Liie factor for bending strength
----
13
8.2.7
Km
Load distribution factor for bending strength
- - - -
11
8.2.1
KR
Reliability factor for bending strength
----
13
8.2.7
%
Size factor for bending strength
----
11
8.2.1
KT
Temperature factor for bending strength
----
13
8.2.7
KY
Dynamic factor for bending strength
----
11
8.2.1
KX
Lengthwise curvature factor for bevel gear
- - - -
11
8.2.1
bending strength M
Lubricant flow rate
Ibmlmin(kg/min)
8
5.1.2
m
Module ( = 25.4/pd )
11 M
8.2.1
m
Gear ratio (never less than 1 .O)
(mm) ----
9
8.1.2
n
Number of standard deviations
----
14
8.2.7
np
Pinion speed
rpm
9
8.1.2
P
Transmitted power
hp NW)
9
8.1.2
AGMA 91 l-A94
Table 1 (concluded) Symbol
4
Name
pd
Diametral pitch ( = 25.4/m )
%
Unite
First iquatior
Reference paragraph
11
8.2.1
Reliabilii constant
in-l ----
I4
8.2.7
sac
Allowable contact stress number
lb/in* (MPa)
18
8.2.8
sat
Allowable bending stress number
lb/in* (MPa)
13
8.2.7
SC
Contact stress number
lb/in* (MPa)
I2
8.2.2
St
Bending stress number
lb/in* (MPa)
II
8.2.1
%vc
Working contact stress number
lb/in* (MPa)
I8
8.2.8
swt
Working bending stress number
lb/in* (MPa)
I3
8.2.7
Till
Inlet oil temperature
OF(“C)
8
5.1.2
Tout
Outlet oil temperature
OF(“C)
8
5.1.2
82
Contact temperature
“F (“C)
I9
8.3.1
tfl
Flash temperature
OF(“C)
I9
8.3.1
tM
Bulk temperature
I9
8.3.1
u
Speed parameter
OF(“C) ----
6
4.2.4
u’
Average rolling speed
inlsec (mm&c)
1
4.2.4
Ve
Entraining velocity
6
4.2.4
W
Load parameter
in/s (m/s) ----
7
4.2.4
W’
Unit tangential load
lb/in (N/mm)
I
4.2.4
WVr
Normal unit load
lb/in (N/mm)
7
4.2.4
wt
Tangential tooth load
I1
8.2.1
Xl-
Load sharing factor
lb 0’4 ----
7
4.2.4
a
Pressure viscosity coefficient
in*/lb (l/MPa)
5
4.2.4
h
Specific film thickness
inlin (mm/mm)
2
4.2.4
CL
Viscosity
reyns (kPa s)
1
4.2.4
PO
Absolute viscosity
6
4.2.4
V
Coefficient of variation or standard deviation
reyns (kPa s) ----
I4
8.2.7
Pn
Normal relative radius of curvature
in (mm)
I
4.2.4
(Ja
Composite surface roughness
tin @ml
2
4.2.4
Ul,(x
Surface roughness of pinion, gear
crin tw)
3
42.4
AGMA 911-A94
4 Design approach
4.1J Maintainability
4.1 Design requirements and goals
Guidelines for field service work, space requirements, and tool limitations must be specified early in the project.
The design procedure begins with a definition of the application, requirements, and goalsforthe project. It is sometimes diiicuft to clearly define all aspects of the project at the start, but a complete tabulation of the following parameters should be made to provide a working definition of the project.
4.1 .I Power/speed and torque/position The complete range of power and speed or torque and position (actuators) must be tabulated including a definition of growth capabili. A duty cycle definition is required for calculation of life. Within these parameters a design point must be selected for sizing purposes.
4.1.2 Gear ratlo and direction of rotation Gear ratio must be specified with an indication of allowable deviation. Input and output directions of rotation are required and are important in selection of the hand of helix or hand of spiral for thrust direction and lubrication considerations.
4.1.3 Life A clear definition of required gear and bearing system life must be provided. Life is defined at a specified survival level.
4.1.4 Weight System weight is criiical in aerospace applications. A value for gear system weight should be specified as dry gearbox weight or gearbox plus lubrication system weight.
4.1.5 Size limita%ons In most applications, gearbox location and maximum envelope will be defined. These details must be made available to the designer.
4.1.6 Reliability Reliability requirements are typically specified in terms of mean time between failure (MTBF). A historical data base of typical component reliability will permit calculation of system reliability. New products are more difficult to characterize. Tech-
nique& quantify reliability levels must be specified for a new gearbox system.
4.1.6 Cost Aerospace gearing is generally more costly than commercial gearing because of the necessary performance, qualii and traceability requirements. Life cycle cost is often established at the start of the project as a goal or as a requirement. Life cycle cost is defined as the total cost of ownership of a system over its operating fife.
4.1.9 Efficiency In most aerospace applications, gearbox efficiency is an important design consideration because it influences system weight and power requirements. Efficiency requirements and goals will provide the designer a clear indication of the project objectives and may affect key decisions in the design process.
4.1.I0 Altitude/attitude requirements Altitude and attitude specifications are required for lubrication system design, since oil pump and oil passage design are dependent on these parameters. In lieu of any specific application data MIL-E-3!593C provides general requirements for aerospace applications.
4.1.I1 Externally generated gearbox loads External loads can be generated by rotor loads, flight maneuvers, gravity and gyroscopic effects, hard or crash landing requirements, or vibration, as applicable. All must be considered in the design of the gearbox housing, mounts and their effects on misalignment of bearings and gears within the gearbox. Typical loads are given in MIL-E-3593C.
4.1.I2 Mount locations Mount locations must be specified to allow design and analysis of the housing and internal structure under external loading conditions. Mount location requirements may also affect maintainability considerations.
4.1 .I3 Loss of lubricant All military and some commercial aircraft have requirements for operation with loss of lubricant , typically specifying a time and power level of operation after loss of lubricant. These require-
5
AGMA 911-A94
ments must be known to allow the design of a suitable lubrication system. 4.1 .I4 Test requirements Test requirements are sometimes different than those used to design the gearbox. If an unusual test is required it can affect the gearbox design. 4.1.15 Noise requirements The recent trend in air vehicle specification has been to require meeting specified internal noise levels in cabin and cockpit. 4.2 identify design criteria It is sometimes difficult to clearly define design objectives or goals of a gearbox or gearset. Proper identification of design criteria requires application of many disciplines such as elastohydrodynamics, involutometry, geometry, stress analysis, system dynamics, materials, kinematics, vibration, heat transfer, processes, manufacturing, economics, etc. Each of the above disciplines requires that design limits be imposed such as: - Stress limits; - Scuffing (scoring); - Minimum oil film thickness; - Type of mounts, deflections and locations; - Weight and Cost: - Vibration; - Noise. The design criteria which have the largest influence on the final configuration are as follows. 4.2.1 Allowable contact stress The tooth contact (Hertz) stress limit depends on the type of application, required service life, properties of materials used, and the shape of the tooth surfaces near the point of contact before the load transfer begins.
as aircraft flap drive systems, winches, and spacecraft robotic manipulator arms. These loads are the highest loads specified for the gears, and are often two to three times higher than the maximum continuous operating loads. This is particularly true for low speed actuator gearing where there are no significant “dynamic” loads. To properly accommodate these conditions, the designer must evaluate the gear design for maximum compressivestresses at the maximum holding loads. Holding loads are usually specified as limit loads, where there may be no permanent deformation or yielding allowed, and ultimate loads, where deformation is allowed but the gears may not fracture. A value of 3.1 times the shear yield strength may be used as the allowable contact stress for most steels. High strength, through hardened stainless gears, may also be utilized where environmental conditions warrant. The surface durability of these gears may be improved, if required, by nitriding. 4.2.2 Allowable bending stress The allowable tooth root bending stress is a function of the hardness and residual stress near the surface of the root fillet and at the core. 4.2.2.1 Power transmission Power transmission gears are usually case hardened by either nitriding or carburizing to obtain adequate high cycle bending and contact fatigue life. A method for calculation of bending stress, along with allowable limits, is given in ANSVAGMA 2001-B88. 4.2.2.2 Actuator gearing
4.2.1.2 Actuator gearing
Gears which are manufactured from high strength through hardening steels (260 ksi and above), and heat treated to through hardness in the Rockwell C 50+ range, have shown higher bending fatigue strength in the lower fatigue cycle range (i.e. less than lo6 tooth bending cycles) than conventional case hardened gears. Thus, a designer seeking optimum minimum weight gearing should consider the actual cycle life imposed prior to making a selection of either case hardened or high strength through hardened gears for a particular application.
Actuator gears are subject to “holding”loads which
Allowable bending fatigue limits are given in ANSI/
are static loads. These loads occur in systems such
AGMA 2001-B88.
4.2.1 .I Power transmission In high pitch linevelocity gearsets, thedistribution of dynamic load is required for accurate determination of tooth contact stress. A method for calculation of contact stresses, along with allowable limits, is given in ANSI/AGMA 2001-888.
6
AGMA 911-A94
d
4.2.3 Surface temperature The mechanism of surface failure due to a sudden temperature rise is one of the major considerations in aircraft gearing. Each oil has a characteristic criiical temperature independent of gear design and operating conditions. Appendix A of ANSVAGMA 2001-888 defines scuffing as related to the instantaneous temperature rise on tooth surfaces caused by frictional heat. The equations which define the surface temperature rise have begun to adapt dynamic conditions and have become more representative of what happens at the gear mesh, including: constrained heat source on the tooth profile, sliding velocity variations, tooth surface conditions, load sharing, oil jet cooling, oil jet impingement depth and air/oil mist cooling. Experiments have verified that minimum values of surface temperature occur at operating pitch diameters. A method of calculating surface temperature is presented in Appendix A of Maximum values ANSVAGMA 2001-B88. generally occur at or near the highest point of single tooth contact. Although the above procedure is currently in use, the method is only applicable under boundary lubrication conditions. Allowable scuffing temperature values should be based on the lubricant temperature at which lubricant breakdown occurs, the material tempering temperature, or the user’s experience whenever possible.
is unit tangential load, lb/in (N/mm).
The currently used lubricant film thickness analysis is the extension of a bearing film thickness study by Osborne Reynolds. Ertel, Gruben, Hamrock, Dowson, and Higginson contributed to the equation in its current form. The most influential parameter in the calculation of film thickness is the speed parameter U, which represents the average rolling speed and the surface condition of the point at which the EHD film thickness is calculated. Surface geometry and finish are important to the EHD lubrication process. EHD theory is based on the assumption of perfectly smooth surfaces, that is, no interaction of surface asperities. In reality, this is not true for boundary lubrication. Therefore, the relative life chart was introduced.
h=+-
. ..(2) u
(31 and 02 are the roughnesses of the two surfaces in contact and his the ratio of EHD film thickness to composite surface roughness. A plot of h vs. relative life is shown in figure 1. This figure assumes sufficient loading and otherwise satisfactory operation of the gears. NOTE-& issupplantingrms as a way of describing roughness. Both terms are still in use but are not equivalent.
4.2.4 Lubricant film thickness Lubricant film thickness has received ever-increasing attention since the time it was introduced by Martin in London Engineering in 1914. H=
. 4896 ’ “h
H = ,,,,g:”
Aerospace gears
//
-i g a, 5
, A Bkaririgs /
” pn W’
where
4
H is oil film thickness, in (mm); p is viscosity, reyns (kPa s); U’ is average rolling speed, inls (mm/s); h
2.2
is normal relative radius of curvature, in (mm):
-6
1
2
ecific film thickness, h
4
h < 0.4 Danger of scuffing for carburized gears li 5 0.4 Acceptable, assuming boundary layer lubrication Figure 1 - Relative life as a functio~~~&n@&
AGMA 911-A94
Further studies by NASA simplified the general equation to the form presented in Appendix A of ANSVAGMA 2001-B88 for dimensionless minimum film thickness: GO54
H mh = 2.65
uo.70 WO.13
.*p-(4)
where the following are dimensionless parameters: materials parameter, G; G= a,?$ . ..(5) speed parameter, V; u=
- PO
42.5 Structural integrity Structural integrity is achieved by the proper definition of gear, bearing and gearbox mounts; gear configuration and materials: selection of bearings; type of bearings and bearing location; seals and type of sealing surfaces. 4.3 Preliminary design
Ve
The areas of concern during the preliminary phase of aerospace gearbox design consist primarily of performance, cost, configuration and packaging.
2w+a
load parameter, W, e(7) where
45.1 Configuration study
a
is pressure-viscosity coefficient, in2/lb (mm*/N);
p.
is absolute viscosity, reyns (kPa s);
Ve
is entraining velocity, in/s (m/s);
E,
is reduced modulus of elasticity, lb/in2 (N/mm2);
p,,
normal relative radiusof curvature, in (mm);
In the preliminary design stage, it is generally necessary to lay out various gearbox configurations which meet the basic speed, power, and ratio requirements. These configurations can be compared against design requirements and rated against each other in terms of reliability, efficiency, maintainability, cost, size, weight, and similarity to past experience. From this process the most suitable configuration for the particular application is selected.
Xr
load sharing factor;
4.3.1.1 Gearing
wr
normal unit load, lb/in (N/mm).
A large number of gearbox configurations are possible to achieve the desired design goal, some of which are described below. The gearbox envelope is generally set by the space available plus the speed, power and ratio requirements. However, the configuration may be further complicated by pitch change mechanisms, accessories, overrunning clutches, engine air intake, etc. Possible configurations include:
The following enhancements may be added to the calculation as follows: - Transient squeeze film effects from change in entrainment velocity, surface geometry and dynamic load; -Actual dynamic&ad profile in place of average tangential load; - Equilibrium surface temperature and oil inlet temperature which defines the temperature of the oil film; - Use of optimal, experimental heat transfer coefficients when oil jet cooling is used for minimization of surface temperature; - Effects of oil entrapment on long face width gears may be included and equations may be separated from short face width gears. The relative film thickness, as calculated using equation 4 for ZYZ~ has been derived and used successfully using narrow face width gears such as
8
accessory gears. If equation 4 is used for power gearing without the previously noted enhancements, the definition of when boundary lubrication occurs may be as low as R = 0.2 to E.= 0.4.
- Offset. This refers to a gearbox in which the input and output shafts have a parallel offset; - Inline. This refers to a gearbox axis in which the input shaft and output shafts are concentric; -Angular. This refers to a gearbox in which the input and output shaft are at an angle to each other. 4.3.1.2 Epicyclic In the same sense that some gearformsare specific cases of a more general configuration (example: A spur gear is the special case of a helical gear with a
AGMA 911-A94
zero helix angle), a gear system can be general or specific. In the context discussed here, we will consider the parallel axis epicyclic rather than the more general bevel epicyclic. Refer to ANSVAGMA Standard 6028-A88 or 6128-A88 Metric. Kinematically, the general case for the parallel axis epicyclic is an arrangement of six gears in two planes as shown in figure 2. By definition a sun gear is a gear element whose axis is coincident with the system axis. Thus, the system shown contains four sun gears; i.e., two external suns and two internal suns. Internal sun gears are sometimes called ring gears. The sun gears of each plane are meshed with an idler. If the two idlers are assumed to be mounted on a common shaft which is, in turn, supported by bearings to a rotatable structure we have the general parallel axis epicyclic system. By controlling the location of the instant center of rotation in the above system of gears, the designer can produce 88 epicyclic variations, each with its own unique properties. Some of the more important variations have been given names and appear in countless transmission systems. For example: - The simple epicyclic: If each of the corresponding gears in the general system are assigned identical tooth counts, then the gearing in one of the planes becomes redundant and may be eliminated, leaving a single external sun, a single internal sun, each meshed with a common idler which is finally supported by the rotatable structure usually called a “carrier’ In the general simple epicyclic, everything in theory can rotate. However, by controlling the location of the instant center of rotation, we can produce some very interesting and important gear systems. These include: - The simple planetary: If we constrain the internal sun against rotation its pitch circle has zero angular velocity and the remaining three components, the external sun, the idler, and the carrier are free to rotate. As the idler rolls in mesh with the fixed internal sun it orbits about the system axis as it rotates about its own axis, thus the idler in a simple planetary has come to be called a “planer. The use of a single planet would place serious balance constraints on the gear system, so it is common practice to fit the carrier with mul-
tiple, equally spaced planets to assure a balanced system, and most importantly, provide multiple load paths for reduced weight.
11
T Figure 2 - The general parallel-axis epicyclic gear train If the input to the simple planetary is to the external sun gear, the resulting gear box will be a speed reducer, and conversely if the input is to the carrier, the resulting gearbox will be a speed increaser. In application the practical usable reduction ratio will lie between 2.5 and 7 and the input and output shafts will have the same direction of rotation. - The star gearbox: If we constrain the carrier against rotation, the system instant center of rotation is coincident with the axis of the idler and the rotating components become the central external sun, the idler, and the internal sun. Since the idler no longer orbiis about the system axis it is usually called a “star”. Again, for reasons of equilibrium and load division it is common practice to fit the stationary carrier with multiple, equally spaced stars. If the input to the star gearbox is to the central external sun, the resulting unit will be a speed reducer, and conversely if the input is to the internal sun, the resulting unit will be a speed increaser. In application the practical usable reduction ratio lies between 1.5 and 6 and the input and output shafts will have opposite directions of rotation. The star gear system has found extensive use in the first reduction of high speed systems because it is
9
AGMA 911-A94
free from high centrifugal bearing loading caused by orbiting planets. - The solar gearbox: If we constrain the external sun against rotation the system instant center of rotation is coincident with the pitch circle of the external sun, and the rotating components become the internal sun gear, the planet, and the carrier. Since, in this system, all components orbit about the central fiied member the name “solar” is quite descriptive. Of the simple epicyclics described so far, the solar system is the least popular since for a given reduction ratio it has higher mesh velocities and a lower transmission efficiency. Usable ratios lie in a narrow band between 1.14 and 1.5 with driving and driven shafts rotating in the same direction. - The compound epicyclic: Referring once again to figure 2, if the tooth counts of the gear elements on each end of the idler shaft are not the same, then all elements in the system can be relevant to the creation of useful gear arrangements. A few of the possible arrangements are noteworthy and will be discussed further: - The compound planetary. If either of the internal suns is constrained against rotation its pitch circle has zero angular velocity and the remaining four components are free to rotate; i.e., the two external suns, the compound planet, and the other internal ring gear. In theory, the designer could produce a transmission with three output shafts, but it would be a rare system where such a configuration would be useful. There are numerous examples of flight systems with counter rotating propellers which use the concept of a compound planetary withtwooutput shafts. As with the simple epicyclics, it is usual practice to configure the gearbox with multiple equally spaced planets to assure a balanced drive, and multiple load paths. In space robotic systems, extensive use is made of the compound planetary using a single driving external sun, one fixed internal sun and one output internal sun. In this latter case, the carrier and the second external sun of the general arrangement are not utilized, and are therefore discarded. Usable ratios available from the compound plane tary cover a very wide range and can be found as low as 5 to values well over 1000. The user is
10
cautioned however, that some compound planetary variations exhibit very poor transmission efficiency due to high effective pitch line velocity in the high torque meshes. A thorough analysis of each application is recommended before committing the design to detailing. 4.3.1.3 The parallel axis differential This special case of the parallel axis epicyclic will be mentioned here because of its extensive use in spacecraft and other systems that require a redundant drive source. In such a system, use is made of two suns, and two planet pairs. Each planet pair is in mesh, and the first planet of each pair is in mesh with one of the sun gears while the second planet of each pair is in mesh with the other sun gear. The carrier is free to rotate and is usually assigned to be the output member. A motor/brake combination is fitted to each of the input suns. In service, either of the motors can be the system input, and the opposite brake can serve as the system reaction member. The reduction ratio of the differential is 2. 4.3.1.4 Accessory drive system The accessory drive system is a drive train dedicated to drive accessories (i.e., lube and scavenge pumps, alternators, generators, etc.) which are requirements of the application. The size and location of the gearbox are dependent on the accessory requirement, positioning of these accessories and the position of the gearbox input drive. When positioning the accessory gearbox, consideration needs to be given to the overall configuration to ensure that a compact package is obtained. Definition of an accessory drive system depends on the spaces and the location available to driie the accessories. One concern is the selection of gear and bearing diameters to fill the distance between the power input and available accessory mount locations. Another concern is to ensure that system life is compatible with the general requirements. Both concerns are equally essential for a successful drive train. Refer to ANSVAGMA 6123-A88 arrangements.
for specific
45.2 Gear sizes There usually are two modes of operation which size gears as follows: - Start up conditions;
AGMA 911-A94
- Spectrum of speed and torque or torque and position operating conditions. Gear tooth geometry definition plays an important role in optimization of design. In general, proper selection of tooth proportion, pressure angle, and spiral angle or helix angle are important to increase the overall contact ratio and to provide better balance between operational stresses. Most recent experience shows increased use of high transverse contact ratio spur and helical gears.
apply to a given material and number of cycles of operation, as shown in the note. Erduy
limit
Case carburized AMS6265 Material
4.3.3 Loads Steady stress
4.3.3.1 Dynamic loads High power density gearing depends on designing gears to maximum load capacity. For high speed gears, a major requirement becomes the ability to accurately calculate dynamic loads so the essential performance and design characteristics can be accurately predicted. Dynamic load consists of three parts. The first part is defined by the component or system resonance. The second part is defined by the gear tooth mesh transmission error, and the third part consists of inputtorquefluctuation. Resonancecondiiions may be controlled by changing gear and web configuration, by damping, or by changing gear ratio. The gear tooth mesh generated conditions can be controlled by changing the tooth form, contact ratio, and damping. Additional mass elastic analysis can also be performed to ascertain torsional dynamics. 4.3.3.2 Centrifugal loads Centrifugal Loads in gearboxes result from component rotation. These loads impose steady stresses, for a given speed, on components such as gears, bearings, and gearbox mounts. The stresses are accounted for under combined loading. 4.3.3.3 Combined loads (Goodman diagram) For high speed gears the steady stress (centrifugal) should be combined with the alternating stress at the gear tooth root, as shown on a Goodman diagram in figure 3. Combined operating stresses such as point A fall within the area bounded by the endurance limit and yield strength and are acceptable. The same gear operating at a higher speed, point B, might fail since the combined stress exceeds the allowable limit. Goodman diagrams
Figure 3 - Goodman diagram for combined loads 4.3.4 Rotating components Preliminary design is not complete unless all rotating components such as splines, shafts, bearings and seals are defined along with the gears. 4.3.4.1 Splines Splines are used to transmit torque between two parts with a common axis. In a gearbox, splines transmit torque between a shaft and a gear or a shaft and another shaft. In practice there are straight sided, serrated, and involute splines but in aerospace transmissions, involute splines are normally used. Involute splines transmit torque through contact between external and internal spline members independent of the fii clearance. This high degree of contact reduces the wear and the length of spline required. The mating internal and external involutes provide a centering effect and the distribution of force from top to bottom is also very good. Requirements for involute splines are usually specified in one of the following standards: -ANSI B92.1, Involute Splines and Inspection; -ANSI B92.2M, Metric Module /nvo/ute Splines and Inspection; - IS0 4156, Straight Cyl~ndricallnvolute Splines. Involute splines can be of the side bearing, major diameter, or minor diameter fii type. In the side bearing fit types of spline, which is the most widely used type, the mating members contact only on the driving sides of the teeth with clearance between the major and minor diameters. In the major diameter fit type of spline, the mating members
11
AGMA 911-A94
contact and are piloted by the major diameters with clearance on the minor diameters. Minor diameter fit splines are only used in situations where the diameter is too small for the cutter of the internal member.
accessory drives are sometimes designed with non-metallic muff inserts between spline members. These serve as an inexpensive compliant part which mitigates metallic spline wear.
The splines can be designed to act as fiied, non-working types or flexible, working types. In the fixed spline, the members are piloted on one or both ends, so that the pilots, rather than the spline teeth, carry any radial load. The fiied type of splined joint is often clamped in the axial direction. The objective in the fiied spline design is to force the spline to carry only torque while other elements carry radial and axial load. Fixed splines must have clearance because of non-concentricity between the spline pitch diameter and the mounting diameters. Without clearance, the internal and external members could bind, leading to increased operating stresses.
Bearings used in aerospace applications generally are one of the following types: - Deep groove ball bearings; - Cylindrical roller bearings; - Needle bearings; - Angular contact ball bearings; - Angular contact ball bearings with split inner race; - Tapered roller bearings; -Journal bearings; - Thrust bearings: - Duplex bearings.
A flexible spline is not held radially by a diametral fit. This permits both radial and angular misalignments of the mating members. There is generally no axial clamping in a flexible spline since this would tend to restrain angular or radial motion. The spline should have enough clearance to allow it to move in a misaligned condition without binding. Splines which must accommodate excessive misalignment should be crowned along the flank to prevent end loading and keep the load toward the center of the tooth. Outside diameter crowning is also used to ensure adequate root clearance under misaligned conditions. A spline subject to angular misalignment carries an induced bending moment across mating members because friction at the spline teeth does not permit free angular motion. The magnitude of the induced moment is a function of torque, coefficient of friction, angular misalignment, and component bending stiffnesses. Lubrication is beneficial to fixed splines and is recommended for flexible splines, especially at high speeds where the teeth tend to have more sliding and wear. Filtered oil supplied to the spline joint provides cooling and also washes away abrasive particles. Grease packed splines are also used. However, they tend to trap the abrasive particles,
4.3.4.2 Bearings
As the bearing size increases, it is generally more difficult to obtain calculated life due to changes in preload caused by mounting, thermal and centrifugal load variations and deflections. 4.3.4.3. Seals The gearbox design is required to minimize the number of static oil or grease seals to prevent lubricant loss. Experience has shown that the use of flat gaskets as static seals has been so poor that they should be used only if absolutely necessary. O-ring seals are generally used. The dynamic seals can either be spring or magnetically loaded face seals, bore rubbing seals, labyrinth seals, or lip seals. Efforts should be made to positively drain, and to provide pressure balance and damping for any dynamic seal system. Consideration should be given to the surface finish and lay of shafts and journals which have contact with seals. Either too fine or too coarse a surface finish could be detrimental. 4.3.5 Lube system requirements Details of the lube systems are discussed in clause 5. Consideration should be given to cool, lubricate and scavenge all rotating power transmission components. 4.4 Detail design
which can accelerate wear and thus will require
Detail design of a geared system requires accurate
periodic maintenance. Flexible splines used as
evaluation of dynamic gear tooth loads caused by
12
AGMA 9ll-A94
load transfer from one mesh to another and momentary overloads caused by system resonance. In detail design, structural gear analysis requires an assessment of tooth load capacity, to select or calculate derating factors. The design process may be based on conventional AGMA or FE analysis. Manufacturing tolerances, tooth errors, profile modifications and system misalignment will significantly influence gear tooth load along the contact path, thus affecting load sharing. Accurate evaluation of gear tooth load sharing behavior under dynamic conditions is not only important in minimizing the weight of the entire system but also is valuable to enhance over- all system reliability. Detail design of aircraft gears can also involve modifications of analysis methods, using nonlinear multibody dynamic analysis including equilibrium analysis, kinematic analysis, vibratory analysis with open loop systems, closed loop systems and elastic (flexible) and/or rigid body systems. All of the above can be used to perform an assessment of the load distribution along the contact line. ANSVAGMA 20014388 defines load distribution for gears of general use. In addition to materials and design configurations, the following items greatly influence the rate of load transfer, or a system’s response to input torque: - Geometry of Pinion and Gear Teeth; - Thermal Distortions; - Gear Rim Centrifugal Forces; - Profile Modifications and Crowning; - Manufacturing and Alignment Errors; - Instantaneous Angular Position of Gears; - Rotational Delay of Driven to Driving Gear (Angular Acceleration); - Total Tooth Deflections (Rim, Web, etc.); - Shaft Deflections (Bearing, Housing, etc.). Load distribution is influenced by the above factors and is non-uniform along the contact lines of meshing gears. To determine tooth load distribution, tooth and rim deflections are required. These deflections vary with the load position and affect the
4.4.1 Finite element modeling considerations Single flank element models can be used to determine tooth stress. To develop a finite element methodology and a design tool to analyze the load sharing behavior from simple spur gear systems to more complex helical and spiral bevel gears on combined systems, an attempt should be made to address the factors influencing load sharing discussed earlier. 4.42 Tooth bending and contact stress considerations Once the load distribution along the contact path is obtained, the calculated load can be transferred to gear tooth pair mesh locations to obtain stresses at the root or along the contact surfaces. The calculations and limits are discussed in clause 8. Gear stresses are a valuable design tool in determining thesize of the gears, and thus minimizing the gear system weight. It is particularly important in sizing (where possible) to base the selection of derating factors of a new design on old designs which are similar and have been successful in the past. The tendency of gear teeth to pit has traditionally been thought of as a surface fatigue problem in which the prime variables are the compressive stress at the surface, the number of repetitions of the load, and the endurance strength of the gear material. In steel gears the surface endurance strength is quite closely related to hardness, so stress, cycles, and hardness become the key items. Gear work in the 1970’s led to two very important conclusions. - Pitting isvery much affected by lubrication conditions; - There is no pitting endurance limit. (S-N diagram does not become asymptotic.) The allowable stress used for design purposes considers such items as the number of cycles and the types of material and oil used. Work on the theory of EHD showed that gears and rolling-element bearings often developed a very thin oil film that tended to separate the two contacting surfaces so that there was little or no
dynamics and tooth root stress as the tooth rotates
metal-to-metal contact.
through the entire mesh.
situation was obtained, the gear or the bearing
When this favorable
13
AGMA Qll-A94
could either carry more load without pitting or run for a longer time without pitting at a given load.
finish, and designed to controlled surface finish and waviness.
Gears in service frequently run for several thousand hours before pitting starts, or becomes serious. A gear can often run for up to a billion (1Og)cycles with little or no pitting, but after 2 or 3 billion (2 or 3 x 1Og) cycles, pitting, and the wear resulting from pitting, can make the gears unfit for further service.
Aircraft bearings are typically AFBMA grade 5 to 7 or better, selectively designed to meet performance requirements.
4.4.3 Regimes of lubrication To handle the problem of EHD lubrication effects, three regimes of lubrication should be considered (see figure 2.12 in [19]*). These are: - Regime I: No appreciable EHD oil film (boundary); - Regime II: Partial EHD oil film (mixed); - Regime Ill: Full EHD oil film (full film). Regime I is encountered in aircraft gears when speeds are jaw, such as in the final stages of gearing in a helicopter gearbox. Regime II is characterized by partial metal to metal contact. The asperities of the tooth surfaces hit each other, but substantial areas are separated by a thin film. Regime II is typical of medium speed gears, highly loaded, running with a relatively thick oil and fairly good surface finish. Most helicopter or final stage turboprop gears are in regime II. In Regime Ill the EHD oil film is thick enough to essentially avoid metal-to-metal contact. Even the asperities generally miss each other. The high speed gear is generally in Regime Ill. In the aerospace gearing field, turboprop drives are high speed and in Regime Ill. Helicopter gears are in the high speed gear region at the input sections of the gearbox. Definition of endurance limits and regime of lubrication are outlined in clauses 5 and 8. 4.4.4 Considerations
for quality levels
Quality levels of aircraft and aerospace gears, bearings and seals are usually as high as system cost limitations permit or as good as can be obtained by using today’s manufacturing methods. Aircraft engine gears are generally ground to obtain quality 12 or better, honed to obtain good surface
High speed aircraft seals are in general carbon face and rotating. Their designs are selected to be flat within two Helium light bands, where each band step measures 11.6 pin (294 pmm). In lower speed applications, lip seals are often used. 4.4.5 Lube system considerations Details of lube systems are discussed in clause 5. Aircraft or aerospace gearbox components rely on direct and pressurized lubrication for the formation of EHD films and cooling. Lube system design includes internal coring or external piping, jets, spray bars, and into mesh or out of mesh lubrication. Lube pumps,deaeration, and filtering requirements are also considered an integral part of the lube and cooling systems. 4.4.6 Tradeoff considerations Completion of final design can also include a comparative study for advanced materials vs. conventional materials. This study includes all rotating components and housings. Life, weight, cost and maintainability can be compared. 4.4.7 Test considerations Completion of any aircraft or aerospace gear system design also includes modification of test tools and test setups to run the following: - Manufacturing Tests; - Component Tests; - Loss of Oil Tests; - Power Plant Tests; - Overload Tests: - Ground Tests; - Flight Tests. These tests are conducted at specified environmental conditions outlined in clause 6. Vibrations, fire resistance, weapons effects, emissions, and attitude are also integral parts of the above defined tests.
* Numbers in brackets[ ] refer to references listed in Annex C.
14
AGMA 91%A94
5 Lubrication 5.1 Cooling vs. lubrication requirements Proper lubrication of gears consists of: a) selecting the correct lubricant; b) ensuring that the lubricant gets into the gear mesh; c) providing adequate lubricant flow so that heat generated in the mesh is removed. There are a number of other considerations in the design of an aerospace gearbox lubrication system but all are related to these three basic requirements. Failure modes that can occur due to inadequate lubrication include: scuffing, micropitting and spalling. 5.1 .l Elastohydrodynamic and lambda ratio
(EHD) lubrication
The thickness of the protective EHD oil film can be calculated using the techniques described in appendix A of ANSVAGMA ZOOl-B88. The ratio of film thickness to composite surface roughness is called the lambda ratio. At a lambda ratio of one, there is theoretically no metal to metal contact. As the lambda ratio decreases, more and more contact occurs. However, carburiied aerospace gears operate successfully at lambda ratios as low as 0.4 without incurring suface damage. Aerospace gears can operate successfully at lambda ratios below 0.4 if adequate boundary lubrication is available. Boundary lubrication utilizes the chemistry of the tooth surfaces, the lubricant and its additives to provide a protective film. Since this type of lubrication is not well understood today, the designer must match the application to past successful1 designs operating under similar condiiions. 5.1.2 Cooling the gear mesh In oil lubricated systems, the amount of lubricant supplied to the gear mesh depends on the heat generation rate. The amount of oil required in the formation of an oil film is miniscule compared to that required for cooling. Most aerospace lubrication systems are designed to handle the highest heat load and have excess capacity at all other operating conditions. Heat generation in gears and bearings can be estimated by various techniques [l] thru [7].
Typically, convection and radiation are ignored such that the entire heat load is to be transferred to the cooling oil by conduction and then removed from the system with a separate oil cooler. When using grease lubrication, solid lubrication and low flow splash lubrication, heat must be removed entirely by conduction through the housing walls or through shafting. Cften cooling is a major limitation of these systems. Knowing the heat load, the lubricant characteristics and the allowable temperature rise, the required oil flow rate can be calculated: HG
= M ch (T,,,,,- Td
. ..(8)
where HG
is heat generated at design point, Btu/min (kJ/min);
M
is lubricant flow rate, Ib/min (kg/min);
ch
is lubricant specific heat at ( Tout+ Th) /2, Btu/lbm”F (kJ/kg”K);
Tau is average oil out temperature, “F (“C);
l-ill
is average oil in temperature, “F (“C).
5.2 Lubricants 52.1 Liquid lubricants Liquid lubrication predominates in the aerospace industry today. Many gear systems must be designed to utilize lubricants that were originally formulated for high temperature turbine engine applications (MIL-L-23699 and MIL-L-7808). In some cases the engine and gearbox use a common lubrication system and thus must utilize engine oil. In other cases a common lubricant has been required to prevent mixing of two diierent types of oil. These lubricants were formulated to meet criteria such as cold flow/cold start requirements, high temperature limitations, material compatability requirements and cost. These properties are derived from fluid base stocks that are not necessarily ideal for lubricated contacts in a gear drive system. Recently a new version of these engine lubricants has been put in service for helicopter applications (DOD-L-85734). This lubricant isvery similarto MIL-L-23699 butaddiiives beneficial to the transmission are included. Tables 2 through 5 list pertinent properties of the most commonly used aircraft lubricants today.
15
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Table 2 - Aerospace lubricant viscosities Tern1 OF
lrature “c
400 204 350 177 320 160 212 100 210 98.9 104 40 100 37.8 -40 -40 -65
ViSCOS
MI L-L-23699’
MIL-L-7808’
1.25 1.63 2.00 5.00 5.ooto5.50* 25.00 25.00 min* 13000max* -
1.00 1.25 1.47 3.00min” -
I,csf DOD-L-85734* 5.00 to 5.50 -
12.00 -
25.00 min c9500 -
2000 13OOOmax
Dexron II3 -
2.23 2.8 7 42 20000 -
dotesReference - AFAPL-TR-71-35; ! QOD-L-85734(AS) specification ) General Motors Dexron II Specification ’from MIL-L-23699D or MIL-L-7808J specifications
1
Table 3 - Aerospace lubricant densities Teml
Densi
OF
392 320 302 212 104 60
200 160 150 100 40 16
MIL-L-23699’
MlL-L-7808’
0.86 0.89 0.90 0.94 0.98 -
0.81 0.84 0.85 0.89 0.93 -
, s/ml DOD-L-85734*
-
0.87
Jotes Reference - AFAPL-TR-71-35 1Exxon Datasheet - ET0 25 I General Motors Dexron II Specification
Table 4 - Aerospace lubricant pressure-viscosity Temperature “F 400 350 320 212 104 l
“C 204 177 160 100 40
Reference AFAPL-TR-75-26
16
Dexron II3
coefficients
Pressure-viscosity coefficient, (in2/lb)x10000[(mm2/N)x10 000] MIL-L-7808G* MIL-L-23699* 0.498 (72.2) 0.428 (62.1) 0.532 (77.2) 0.462 (67.0) 0.556(80.6) 0.486 (70.5) 0.681 (98.8) 0.613(88.9) 0.966(140.1) 0.918 (133.2)
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5.2.2 Greases Greases are commonly used to lubricate actuator gearing and gearbox components such as bearings and splines. Several grease lubricated helicopter transmissions are in production but are not common. The most common greasesand their uses are listed in table 6.
Table 5 - Aerospace lubricant specific heat values Temperature
“F
Specific heat, btu’lb OF[kJ/(kg OK)] MIL-L-23699 or MIL-L-7808G*
“C
5.2.3 Dry lubricants Dry lubricants are widely used in spacecraft systems where liquids or greases cannot be used due to out-gassing problems (see clause 13) and also in aircraft systems where liquids or greases cannot be contained. These lubricants do not provide the same level of protection as liquids or greases. Thus, the applied loads and sliding velocities must be significantly lower in these systems. Table 7 lists common dry lubricants in aircraft use today.
t
400 350 300 250 200 150 100
0.562 0.551 0.538 0.524 0.508 0.486 0.464
204 177 149 121 93 66 38
(2.35) (2.31) (2.25) (2.19) (2.13) (2.03) (1.94)
’Reference AFAPLR-T&75-26
Table 6 - Aerospace greases* ML Specification
Description
Application
MIL-G-6032 MIL-G-21164 MIL-G-23827 MIL-G-25013 MlL-C-38220 MIL-L-27617 MlL-G-46006 MIL-G-81322 MIL-L-81827 MIL-G-83261 MIL-G-83363
Oil resistant grease Molybdenum disulfide Gear and actuator grease Aircraft bearing grease High speed bearing grease Oil resistant grease Aircraft grease General purpose grease High load capacity grease Extreme pressure grease Helicopter transmission grease
Tapered plug valves, gaskets Splines, sliding steel surfaces Bearings, gears, etc. Ball and roller bearings to 200 000 DN Ball and roller bearings to 400 000 DN Tapered plug valves and gaskets Driveshaft couplings Bearings and gearboxes Splines and bearings Gearboxes, actuators Tail rotor and intermediate gearboxes
Military Handbook, Guide for Selection of Lubricant and Compounds for Use in Flight Vehicles and Components, MIL-HDBK-2754, May, 1969 NOTE -The above greases are not to be used in vacuum applications(see clause 13).
l
Table 7 - Aircraft dry lubricants* MIL Specification
Description
Application
m-G-659 MIL-M-7866 MIL-L-8937 MIL-L-23398
Graphite Molybdenum disulfide Corrosion inhibiting Air drying solid film
Dry lubricant or mix with oil Threads, gears Gears, flap hinges Steel, titanium, aluminum
l
Military Handbook, Guide for Selection of Lubricant and Con-pounds for Use in Fli&ht Vehicles and Components, MIL-HDBK-275A, May, 1969
17
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5.3 Distribution systems in this subclause, guidelines are described for selection of various gearbox lubrication schemes. The development of any new gearbox must, however, include testing to determine the success of the lubrication system design. There are many factors that can unexpectedly affect system performance. A successful lubrication system design prevents gear tooth scuffing, meets oil temperature rise expectations and limits power consumption due to oil distribution and removal. 53.1 Pressure-fed A pressure fed system, for this discussion, consists of an oil supply line directed at a gear mesh which is fitted with a metering orifice. 5.3.1 .l Oil pressure requirements Oil pressure is required to propel lubricant into a rapidly moving ,gear mesh. An oriiice is used to convert oil pressure into kinetic energy creating the oil velocity required to penetrate the mesh. By properly selecting both the oil pressure and the orifice (jet) diameter, the flow rate and oil penetration depth can be controlled. 53.12 Analysis for pitch line velocities to approximately 25 000 feet per minute (127 mkec) For gear meshes operating at 25 000 ft/min or less the technique described in [8] can be used. Oil penetration depth into the mesh is simply a time of flight calculation. At these speeds the jet still acts as a solid stream and is not significantly affected by air turbulence. Ati long as the oil flow is not interrupted by the passage of another tooth, oil will penetrate to the depth as determined by this method. 5.3.1.3 Analysis for pitch line velocities greater than 25 000 feet per minute (127 m&c) Standard analytical methods have not been confirmed above 25 000 feet per minute. Windage of rotating parts must be considered to ensure proper targeting of the oil jet. Development testing of the unit is required to determine the adequacy of the lubrication scheme employed. 5.3.1.4 Oil jet orientation Oil isfrequently supplied toward thedisengagement side of the mesh on the driving side of the tooth. Power losses are minimized, cooling oil will reach
cooling is maximized. In some cases only a small amount of oil is required to provide adequate into mesh lubrication can increase cooling. churning losses due to the presence of excess oil in the mesh. In high power applications it may be necessary to use multiple oil jets to provide the required oil flow and still maintain jet velocity. Oil jets can be placed on both the out of mesh and into mesh sides to obtain the required oil flow rates. In this case the intomesh jet should supply a lesser flow. 5.3.2 Splash Splash lubrication is used when heat generation rates are low compared to cooling available either through conduction through the gear or through convection to a nearby lubricant. A first order heat transfer analysis can be performed to determine the cooling available from the gear and the lubricant. Splash lubrication can be provided by dipping the gear through a pool of oil or by placing the gear in the path of a stream of oil exiting from another mesh. Development testing is typically performed to determine adequacy of the lubrication scheme. Consideration should be given to pre-lubing gearboxes of this type to ensure that oil is available during startup. 5.3.3 Mist Mist lubrication is similar to splash lubrication except that the cooling provided by the oil is very low. The oil available in the mist is sufficient only to allow generation of an oil film. In some cases an air/oil mist is created with the intent that the air will provide cooling of the mesh. 5.3.4 Grease Grease is used for gearbox lubrication when a sealed area can be provided to prevent contamination and lubricant loss. The advantages include low cost and simplicity particularly when it would be diicult to introduce a liquid lubricant system. Heat generation in these components must be low compared to the heat sink available through conduction to the housing walls and the shafting to prevent high operating temperatures. Sealed, grease lubricated gearboxes have been
the tooth flank at its highest temperature thus
used in helicopter tail rotor and intermediate gear-
promoting heat transfer and the length of time for
box applications.
18
In addition to the advantages
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stated above for bearings and spiines they are also less vulnerable in combat situations. Flap actuator gear drives are another common aerospace application for grease lubrication.
transmission and also has less tendency to chum the oil. Wet sumps may not be practical for some applications due to the increased frontal area required or other envelope limitations.
5.3.5 Powder lubrication
5.4.3 Oil deaeration
Powder lubrication is being investigated for very high temperature applications where only dry lubricants can survive. in these systems dry lubricants are blown into bearings and gears as an aerosol to lubricate and cool the system.
Oil foaming is a major concern in aircraft gear systems. Air is easily trapped in oil due to the mixing action that occurs during high speed rotation of gears and bearings. Foaming can be controlled by allowing air to escape naturally in an oil tank, providing deaeration trays in the oil tankor by using air/oil separators. Oil circulation rates must be selected to provide time for air to escape in an oil tank or tray. Since this usually requires a larger oil tank it is sometimes necessary to supplement the natural deaeration with an air/oil separator.
5.4 Lubrication system design considerations 5.4.1 Common vs. separate lubrication systems In some cases the designer may have the choice between a self contained lubrication system or an external system. Frequently, however, the gearbox lubrication system must becommon with the engine or other equipment and the resulting complexities must be considered. Table 8 lists some of the advantages and disadvantages of each type of system. 5.4.2 Dry sump vs. wet sump in a dry sump system, oil is stored in a separate oil tank, not part of the transmission housing. In operation, oil supplied to transmission components drains to scavenge ports where the air/oil mixture is pumped out to the tank. The wet sump system integrates the oil tank into the transmission housing, typically at the bottom. Scavenge pumps are required only for areas that are diiicult to drain. Otherwise, gravity is used to return oil to the tank area. Oil pump inlets are designed to remain covered with oil during all attitudes of operaion. The wet sump system offers the advantages of no external plumbing for connecting the oil tank to the
5.4.4 Oii scavenging systems and oil baffles In order to prevent heat generation due to churning of excess oil near rotating bearings, gears or seals, it is frequently necessary to use oil scavenge pumps to return oil to the sump or tank. If this excess oil is not removed several problems can arise: excessive power consumption due to oil churning, high oil temperatures, oil foaming, oil leaks through seals, and larger oil tank capacity requirement due to change in level on start-up. In the design of the scavenge system the designer must consider attitude and altitude requirements of the system. The scavenging action must continue through all aircraft maneuvers and thus through a range of system attitude alignments. This requirement frequently results in multiple oil scavenge pumps with oil pick-up passages placed in strategic locations. It is generally not feasible to use one pump with multiple scavenge pick-up points since only air will be removed if any one passage becomes exposed to air.
Table 8 - Advantages and disadvantages of a common engine and transmission lubrication system Common system
Separate systems
Lower weight - common parts Engine oil must be used Transmission failure can affect engine
Increased weight - redundant parts Transmission oil can be used Transmission failure is self contained
External plumbing required
Self containedplumbing
19
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Scavenge oil pumps are typically rated at two to three (and sometimes more) times the supply pump flow rate. This ensures that oil will not build up in critical areas under adverse operating conditions. One of these conditions occurs at altitudes where air is less dense and is easily dissolved into the oil. Removal of this foamy m’tiure requires additional flow capacity since the volume of this mixture is greater than solid oil. Design and placement of oil baffles can also affect oil scavenging and gearbox performance. The purpose of an oil baffle is to prevent oil from becoming entrained in rotating gears and bearings and to remove cooling oil that has done its job. Many of the same problems discussed above will occur if oil is allowed to flow into rotating parts. An oil baffle diverts oil toward the sump or scavenge port to prevent a build-up of oil within the housing. A baffle can also be used to control air windage and its effect on oil foaming. A baffle can be cast into the housing or fabricated separately and bolted into position. The baffle should be tested to determine its natural frequency to ensure that it will not be subject to high cycle fatigue in the high vibration environment inside the gearbox. 5.4.5 Pressure drop in oil passages Consideration should be given to pressure drop in oil passages, particularly if the length of the passage is long, if there are a number of sharp turns or if the surface roughness of the passage is poor as in cast passages. If this is not done, the calculated flow rate could be significantly less in operation. A rule of thumb used in the past for sizing passages has been to design for a velocity not to exceed 15 ft/sec (4.6 rn/sec) on the pressure side and 5 ft/sec (1.5 m/se@ on the suction side. 5.4.6 Oil sampling If oil sampling capability is required it must be designed into the system at an early stage. Frequently this feature is added to an existing system and good results are not possible. In order to obtain a representative oil sample, the oil tap should be located at the end of a line just before a turn is made. When the sample is taken, oil from the main stream is obtained. In many cases an oil tap is placed in a region where debris is allowed to build up and the sample is artificially contaminated.
20
When taking an oil sample it is good practice to clean the tap priorto taking the sample and to let the oil flow for several minutes before taking the actual sample. 5.4.7 Fill/drain considerations It is often possible to introduce more contaminants intoagearsystem during top-off than during normal system operation. This is particularly true for systems with fine filtration. A fine filtration system cleanses the entire system over a period of time to very low contaminant levels. Opening the system during top-off may allow debris to fall into the system. Control of the top-off function can partially be aided by design and placement of the fill mechanism but is primarily controlled by training maintenance personnel. Oil changes are common in systems with coarse filtration. Change intervals are determined from past experience with similar systems. Oil changes in systems with fine filtration are less frequent. The cleanliness levels found in fine filtration systems are maintained continuously. The lack of circulating debris tends to reduce the amount of new debris that is formed. In systems in which the engine and gearbox are common, oil consumption by the engine requires a routine tovff of new oil. This replenishes any additives that may have been lost. 5.4.8 Vulnerability/safety Some gearbox specifications include the requirement for an emergency lubrication system independent of the primary system. The emergency system allows safe operation of the aircraft for a limited time in the event of the loss of the primary system. The degree of sophistication in these systems can range from a complete secondary pumping system to a very simple gravity, In addition, some military oil drip system. transmissionsare required to operate for a specified period of time after loss of the lubrication system. 5.4.9 Low temperature operation Most aircraft gearboxes must be designed for operation at a very low temperature, typically4 OF (-54 “C). Since lubricant viscosities are quite high at this temperature, it is diiicult to circulate oil in the lubrication system. The requirement led to the development of the MIL-L-7808 lubricant with its lower cold temperature viscosity. Even with this
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lubricant, oil heaters may be required to aid starting. It is difficult to pump lubricant during start-up at low temperature.
points such as a main distribution line leading to gear mesh oil jets. Screens are also used to protect scavenge pumps.
5.4.10 Oil coolers
5.5.3 Filter rating nomenclature
Some type of oil cooling system is required for most aerospace gear systems. Convective heat loss through the housing may not remove enough heat under all flight conditions to maintain acceptable oil temperatures. Typical oil cooling systems include:
A typical filter rating might be: 10 micron absolute, Beta=1 00. This terminology means:
- ram air/oil heat exchangers; - air/oil heat exchanger with blowers; - fueVoil heat exchangers for lower heat load systems. Oil temperature is sometimes controlled in these systems by varying the amount of air or fuel supplied to the coolers. A typical aerospace lube system schematic is shown in figure 4. Some gearboxes that have low power consumption can be operated without oil coolers if the heat loss is sufficient to keep operating temperatures within acceptable ranges. 5.5 Filtration 5.5.1 Filter types There are three types of aircraft filters in use today characterized by the media used to trap debris: steel mesh, paper and fiber mesh. Steel mesh filters are being utilized less due to: -the maintenance required to clean them; -the loss in effectiveness after cleaning; - cost.
Paper media filters are generally used for coarse filtration levels. Throw-away fibrous media filters are becoming more common as fine filtration is becoming more popular. 5.5.2 Screens Screens or last-chance-filters are used to protect system components from large pieces of debris, The mesh is usually coarse and is intended to stop debris from a failure that occurs upstream of the screen but downstream of the main filter. Screens are typically placed ahead of critical lubrication
The fitter will reduce the number of particles greater than or equal to 10 microns in size entering the filter by a factor of 100. This equates to an efficiency of 99 percent. Or, 1 percent of the particles greater than 10 microns will pass through the filter. Nominal filter ratings are also used today but are less stringent. Nominal efficiencies are typically 90 to 98 percent. 5.5.4 Fine filtration Pine fittration is becoming more common today for aircraft engine and gearbox systems for several reasons: longer bearing and gear life due to lower levels of debris, fewer or no oil changes and potentially longer filter life. Fine filtration requires a 3 to 5 micron absolute rating with a Beta factor of at least 100. Fitter life has been found to increase after the initial system clean-up phase. Once the system is clean, it tends to stay clean. 5.5.5 Filter location Filters can be located on the pressure side of the system, the scavenge side or both. Selection of the site depends on the designer’s philosophy, e.g. a pressure side filter will always protect the component, whereas a scavenge filter will protect the heat exchanger, sump or tank, and will not proliferate the debris. The overall system design must be considered to reach the best decision on filter location. 5.6 Oil pumps The three most common aerospace gearbox oil pumps are the gear pump, the vane pump and the gerotor pump . All are positive displacement pumps since output flow is a function of fifed internal geometry. Flow increases linearly with shaft speed. However, the vane pump can be designed to provide a variable flow rate. 21
AGMA 91%A94
Oil tank
From nacelle
Prjp gearbox lube and scavenge pu$ps To dacelle
+
Gear mesh I I i
Gear mesh Bearings
c/l
kf Gear mesh
Bearings
I
)-I Bearings b
Spline
1 I
Prop gearbox sump
H
Spline
Accessory gearbox sump
LEGEND
0
Screen mesh
0,
Regulating valve
Filter
*
By-pass valve
Figure 4 - Typical aerospace lubrication system schematic 56.1 Gear pumps A gear pump moves a fixed volume of oil from the inlet side to the outlet as shown in figure 5. Plow volume is a function of tooth depth and width. As resistance to flow increases at the outlet, the power
gear versions of this type of pump can increase the flow with little increase in pump size.
5.62 Vane pumps Vane pumps consist of a rotor mounted off center in a circular cavity (see figure 6). Vanes in the rotor are
required to rotate the input shaft increases and
forced out due to centrifugal force under rotation
discharge pressure increases accordingly. Three
and contact the cavity walls.
22
The amount of
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eccentricity determines the volume of flow per revolution. Variable flow rates are achieved by allowing adjustment of the eccentricity during operation. Variable displacement pumps are more efficient since excess flow (and thus excess power loss) is eliminated.
r
Figure 5 - Spur gear pump 5.6.3 Gerotor pumps
Housing body/cam ring r
Rotor
Figure 6 - Vane pump
Gerotor pump action is shown in figure 7. Oil is drawn into an area equivalent to one tooth space at position 1 and is rotated and then forced out due to compression as it moves toward the exit at position 4. Both rotors rotate but the inner rotor driies the w-w Spring type pressure regulators are often used to maintain a constant oil pressure under varying conditions of operation, e.g. speed and temperature. These regulators are commonly built into the pump body. Oil flow rates are then controlled by selecting orifice (jet) sizes to supply the correct oil flow at the regulated oil pressure.
5.7 Lube system condition monitoring 5.7.1 Chip detectors Magnetic chip detectors are used to collect debris circulating in the lubrication system. In some systems the pilot receives an indication when the number of particles collected is sufficient to trigger a signal in the detection circuit. In other systems the chip detector is visually checked as a maintenance function. Chip detector operation is very sensitive to systemconfiguration and sensor location. Testing of chip migration in the housing is recommended for placement of chip detectors.
Figure 7 - Gerotor Pump 23
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There are many types of chip detectors but their function is similar to that just described. The burnoff chip detector was developed to reduce the incidence of nuisance indications during flight. This device automatically passes a strong current through the collected chips, melts them, and may provide an indication to the pilot. This will clear the indicator. If a true failure has occurred chips will collect again causing another indication shortly thereafter. Another type of chip detection system under development has the ability to count particles of several sizes and track the build-up of debris. If the rate of build-up is rapid, there is a good chance that a failure has occurred. 5.7.2 Temperature/Pressure sensors Lubricant temperature and pressure are the primary parameters used to track system condition. These sensors must be placed in the system correctly to obtain good results. As an example, a temperature sensor that is to detect the temperature of the oil exiting the gearbox must be placed in an area where good lubricant flow is taking place. Otherwise, it may just indicate air temperature. 5.7.3 Spectrometric oil programs Spectrometric oil analysis programs have been used for many years to detect wear of components in aircraft systems. The procedure consists of taking oil samples at regular intervals and analyzing them for metal content in a laboratory. These programs have shown varying degrees of success. Due to the size of the particles analyzed, this method may not work in fine-filtration systems. 6 Environmental
issues
This clause illuminates the environmental issues likely to be encountered by an aerospace gearbox during its lifetime, including storage, transport, and operation. Data on the occurrence of environmental extremes may be found in several publications. MILSTD-210, Climatic information to Determine Design and Test Requirements for Military Systems and Equipment provides data on free air (i.e. ambient) conditions. During the development phase of a gearbox, MIL-STD-810, Environmental Test Methods and Engineering Guidelines, can be used to
MlL-STD-462, Measurement of Electromagnetic Interference Characteristics. Another source of environmental information is the NASA Technical Memorandum 82473, Terrestrial Environment (Climatic) Criteria Guidelines for Use in Aerospace Vehicle Development. 6.1 Ambient temperature effects The operating ambient temperature range for the gearbox will be determined by the aerospace system application operating envelope in conjunction with data presented in the International Civil Aeronautical Organization (ICAO) Standard Atmosphere. [Note that the ICAO Standard Atmosphere is identical to the US Standard Atmosphere for altitudes below 65 617 feet (20 000 meters) .] The gearbox should be capable of operation at continuous rated power or less at any ambient temperature within the specified operating envelope. It is important to note that at temperature extremes material characteristics and properties can be significantly affected. For example, materials with dissimilar coefficients of thermal expansion (e.g., aluminum versus steel) will significantly influence operating fiis and clearances within the gearbox and need to be adequately considered in the design. Also, at very low temperatures lubricant viscosity will be orders of magnitude higher than at typical operating temperature. As a result, lubricant flow through heat exchangers may not occur even after the remainder of the gearbox is at normal operating temperatures. The net consequence is that power loss through the gearbox will be significantly increased. In addition, increased brittleness and notch sensitivii can become an issue at very low temperatures. For example, the fracture toughness of gear materials can be reduced by more than twenty fiie percent when going from room temperature to -65°F (-54%). Other potential low temperature problems are loss of O-ring resilience, resulting in leakage of lubricant from the gearbox and contraction of the gearbox at a greater rate than the gears and bearings, causing abnormal loading ofthesecomponents. Refertoclause13forvacuum environment applications At the other extreme, lubricant viscosity will be very low leading to the possibilii of very thin oil films in the gear meshes and in the internal bearing con-
gain additionalinsights. Electromagnetic interfer-
tacts. If the oil film thicknessis too small,allowable
ence test and analysis requirements are covered in
compressive stress limits will be reduced for the
24
AGMA 911-A94
gears and bearings and the probabilii of gear tooth scuffing (scoring) or bearing surface distress will be significantly increased. The judicious selection of oils with appropriate properties at both extremes (see clause 5 for detailed information) can help avoid these potential problems as can the proper design and sizing of oil coolers to maintain a desired oil operating temperature. Another consideration associated with very high ambient temperature operation is the retention of adequate surface hardness for both gears and bearings. Sufficient loss of hardness (e.g., two points below design minimum values) can radically degrade compressive fatigue life. Therefore, retention of adequate hardness is a very important consideration in the proper selection of gear and bearing material.
6.3 Attitude effects The gearbox operating attitude limits will vary greatly depending upon the particular aerospace application. For reference, figure 8 shows typical limits for military aircraft. The aircraft/gearbox should be capable of continuous operation in the unshaded area and capable of some period of transient operation in the shaded area. Sometimes diierent duration requirements are specified depending upon whether the system is at zero g (acceleration due to gravity) or negative g operating conditions. At any attitude, the gearbox should be capable of operation at continuous rated power or less. In addition to operating requirements, the gearbox should be capable of being stowed or transported across the specified attitude spectrum. * Above horizontal
Special mention is madeforaerospace applications required to operate at cryogenic temperature (i.e., at liquified gaseous conditions). At these extremely low temperatures, very specialized design practices, materials, and lubricants are needed to ensure proper gearbox operation. 6.2 Ambient pressure effects The operating ambient pressure range for the gearbox will be determined by the operating envelope of the particular aerospace application. Typical factors affecting pressure include operating altitude of the aircraft, rates of climb/dive, and the effects of nuclear or conventional weapons. At extremes, pressure changes, at a rate of 1000 psilsec (6.895 x 1O6 Pa/se@ with magnitudes of 2.5 psi (1.724 x 1O4Pa) and 70 psilsec (4.826 x lo5 Pa/set) with magnitudes of 5 psi (3.447 x 1O4Pa) can be anticipated. The gearbox should be capable of operation at continuous rated power or less at any ambient pressure or pressure change within the specified operating envelope. Design of the gearbox should include consideration of the possible pressure changes on performance. Pressure differentials across gearbox seals can fluctuate and even reverse in flight if not properly designed for and cause seal contamination and oil leakage. Likewise, venting of the lube system is often accomplished within the gearbox through breather pressurizing valves which need to be designed for all possible pressure variations. Adverse consequences include loss of oil, oil entrapment, and oil foaming, all of which can lead to early failure of the gearbox components.
4,
,Roll &-rgle
P
A *Below A
horizontal * Reference to ground n - Test points Notes 1. The gearbox shall be capable of operating at all possible conditions; however, for the purpose of defining the direction of acceleration vector from the engine CG, the figure assumes no acceleration other than grav.w. 2. Gearbox centerline perpendicular to plane of paper. 3. Continuous operation in clear area. 4. Thirty second operation in shaded area.
Figure 8 - Typical gearbox attitude limits
The requirements for variable attitude limits derive from the fact that the aircraft application is required
25
AGMA 911-A94
to climband bank, and formilitaryaircraft, tosustain air combat maneuvers and terrain following operation and possibly undergo inverted flight. If the oil system for the gearbox is not properly designed to operate within the specified attitude limits, serious consequences for the gearbox can result. For example, extended operation in zero or negative g conditions can result in oil system malfunctions which can effect the gearbox such as oil starvation, gearbox flooding, oil foaming problems, and seal leaks. Internal pressurization techniques are often used as solutions to these problems. Special provisions, such as an auxiliary lube system, may be necessary in the gearbox lubrication system for extended inverted flight operation. Placement of suction/feed points to the pumps can alter the attitude envelope. See 5.4.4. 6.4 Contaminant effects (water, corrosives, dirt, dust, and sand) Due to worldwide all weather operation requirements for many aerospace systems, gearbox operation in the presence of contamination has become a necessity. For example, aircraft operating out of airports near arid parts of the world often takeoff and land in conditions where the air contains sand, dust, and dirt. Water gets ingested during rainstorms. The fact that the lubrication system which services the gearbox has to breathe means that atmospheric water and vapor is ingested into the gearbox. This moisture may or may not contain salt which presents a whole host of corrosion related problems. The incidence of contamination is so prevalent that specifications have been developed for them. The table below shows the data for US Specification Sand. The specified sand should include 90 percent silicon dioxide (SiO;z)of angular structure with total particle size distribution by weight as shown in table 9. Table 9 - Particle size distribution by weight Size, fin
Size, CPM
o-2 953 2 953-4 921 4 921-7 874 7 874-15 748 15748-35433 35 433-39 370
(o-75) (75-l 25) (125-200) (200-400) (400-900) (900-l 000)
Percent by weight 5+2 15+2 28f2 36+2 3.5 lk 0.5 1.5kO.5
Using the specification contaminant at some predetermined concentration level, qualification tests are
26
conducted to validate the adequacy of the gearbox or other component design to operate for some extended period of time. For the required duration the gearbox should be capable of operation without any degradation in performance. Obviously, the best wayto handle the contaminant is to preclude its introduction into the gearbox by prudent design. For the larger contaminants, like sand, this may be possible. However, for water or dirt this is probably not possible and other solutions have to be found to accommodate them such as filters to remove the dirt after it has entered or protective coatings to resist the deleterious effects of water. 6.5 Vibration/Shock effects This subclause covers vibration and shock effects on the gearbox ensuing primarily from external sources. Those initiated internally within the gearbox are covered extensively in clause 7. The random vibration environment which an aerospace system gearbox might be subject to can originate from the following sources: - Turbulent aerodynamic airflow along external surfaces of the airframe structure; - Engine noise impinging on the airframe structure; - Rotor/propeller blade induced effects; - Airframe structural motions causing sympathetic gearbox response; - General aircraft motions caused by such factors as runway roughness, landing, and gusts; - General motions encountered during transport of the gearbox, such as by rail, ship, or truck; - For military systems, the gun blast pressure impinging on the aircraft structure from high speed repetitive firing of installed guns and vibration from antiaircraft fire (flak). Typical vibration spectra tested for in aircraft are shown below in table IO and figure 9 for propeller aircraft and turbine engine equipment, and in figure 10 and table 11 for helicopters. The shock environment is often a consequence of the following effects: aircraft launch/catapult and landing; aircraft crash; handling, such as dropping of the gearbox; and transport of the gearbox. Typical shock pulse levels and durations tested for are shown in figure 11. Generally, specific requirements are specified by the customer. These may diier from those in the figure.
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Table 10 - Suggested functional test levels for propeller aircraft and turbine eng,ine equipment Equipment location*- 3
Vibration level (Ll) at F1 4$5, g*/Hz
!
0.1 0.3 0.6 1.0
In fuselage or wing forward of propeller In fuselage or wing aft of propeller In engine compartment or pylons Equipment mounted directly on aircraft engines
1 Fl= fundamental excitation frequency; Fi= source frequency (i = 1 -4), FZ = 2 Fl, F3= 3 FI, F4= 4 F1 2 When panels and racks are not available for equipment installed on vibration isolated panels or racks, or when the equipment is tested with isolators removed, use “fuselage or wing forward of propeller” category with levels reduced 4 dB. 3 Increase test levels 6 dB for equipment mounted on fuselage or wing skin within one propeller blade radius of the plane of the propeller disc. For all other skin mounted equipment, increase levels by 3 dB. 4 Bandwidth vibration around each Fiwill equal f 5% F fpr constant-speed excitation. When excitation is not constant-speed, bandwidth will encompass operating speeds for cruise and high power operation. 5 Fl= 68 Hz for most C-l 30 aircraft.
3
Li
4. I ! 1
-6 dB/octave octave = 2 to 1 frequency range
-
m s 2 45 E 5 Q 6 -01 n
I 15
I
I I
F,
l .
Li
I l
.I I .
I
. .
1 .03
I
h-l
I
F2
F3
I 4
_---_-_ __ 8 I II I -N4
I 2000
I
i5
200;
F,
Frequency (Hz)
Frequency (Hz)
B. Turbine engine equipment spectrum
A. Propeller aircraft spectrum
Figure 9 - Suggested vibration spectra for propeller aircraft and turbine engine equipment
A A i 1 rl :I’
&I57 g-g ES *; $02 i5gj 5m 8a at % %
22
Note: See table 11 for L’t , Lf2, L13and L/4 g-levels.
L’3 L’4
I
t 8 I
8 I I
I
10
F,
F2
F3
I
F4
1 t t I
300
8 8
Equipment location General instrument panel External stores On/near drive train elements
w, , g*A-lz
Ft , Hz
0.002 0.002 0.002 0.02
500 500 500 2000
t
5
Frequency (Hz)
Figure 10 - Suggested vibration spectrum for equipment installed on helicopters 27
AGMA 91%A94
Table 11 - Suggested functional test peak levels for equipment installed on helicopters Equipment location
Source frequency (5) range, Hz Peak vibration level at F,, L’,, g’s
General’
5-25 25-40 40-50 50-500
0.1 F, 2.5 6.5 - 0.1 F, 1.5
5-25 25-40 40-50 m-500
0.07 F, 1.75 4.55 - 0.07 F, 1.05
External stores’
5-25 25-40 40-50 50400
0.15 Fx 3.75 9.75 - 0.15 F, 2.25
On/near drive train elements*
5-50 50-2000
0.1 F, 5 + 0.01 F,
Instrument
panel’
1
F,= Source frequency of interest = Fr, Fe, FJ, or F.4 Fj = Fundamental source frequency Fp = 2F1 ; F3=3Fj ; F4 = 4F1 Upon determining values of Fl, F2, F3, or F4 (figure lo), select the appropriate source frequency range for each when determining peak vibration levels. The source frequency ranges are not presented in order of FI = F4. 2 Ft, F2, F3 and F4 must be determined from drive train areas for the particular helicopter. Note (1) is then applicable.
[[
Ideal sawtooth pulse Tolerance limits
Test Flight vehicle Operation test
Crash safety
Minimum peak value, (P) g’s Ground equipment
Flight vehicle equipment’
402 75
20 40
Note - The oscillogram shall include a time about 30 long with a pulse located approximately in the center. The peak acceleration magnitude of the sawtooth pulse is P and its duration is D. The measured acceleration ulse shall be contained between the broken line boundaries and the measured velocity change (which may be obtained by integration of the acceleration pulse) shall be within the limits of K fO.l I$ , where X is the velocity change associated with the ideal pulse which equals 0.5DR The integration to determine velocity change shall extend from 0.4 D before the pulse to 0.1 D after the pulse.
Nominal duration, (0) ms Flight vehicle Ground equipment’ equipment 11
11
6
11
1 Shock parameters recommended for equipment not shock-mounted and weighing less than 300 pounds (136 kg). 2 Equipment mounted only in trucks and semitrailers may use a 20 g peak value.
Figure 11 -Terminal-peak sawtooth shock pulse configuration and its tolerance limits (for use when shock response spectrum analysis capability is not available)
28
AGMA Qil-AQ4
Some potential adverse effects of shock and vibration on a gearbox include cracking of gearbox housings and covers, breaking of gearbox mounts, cracking of oil supply and scavenge lines, misalignment of internal components, and brinelling/ false brinelling of the gearbox bearings and gears. Several means are typically employed to minimize these adverse effects. For example, vibration damping coatings can be applied to gearbox housings, elastomeric dampers are used for gearbox mounting, pneumatic or hydraulic mounts are employed during transport, and oil lines are amply supported to preclude the presence of potentially harmful resonant vibration frequencies. Qualification tests should be conducted on the gearbox at the appropriate vibration and shock levels and durations (again depending on the particular aircraft application) to validate the gearbox design. Because of potential airframe effects, testing of the total aerospace system is necessary for final evaluation. 6.6 Fire resistance requirements The capability of the aircraft and its many components including the gearbox to have adequate fire resistance is critical to overall aerospace system operational safety. It is essential that any system component containing flammable fluids, such as the gearbox with its oil lubrication system and fuel-oil heat exchangers, will not contribute to an aircraft fire and will allow sufficient time to isolate the fire zone. Gearbox seals, covers, and ftiings are primary areas of concern for leakage and potential fire initiation sources.
resistance requirements. These tests are conducted while conveying the oil within the gearbox at the lowest oil flow rate, highest ambient system pressure, and the highest oil temperature expected over the complete gearbox operating range. 6.7 Electromagnetic effects Electromagnetic compatibility between any gearbox electrical/electronic component and those of the rest of the aerospace system is essential for For example, proper and safe operation. electromagnetic interference could conceivably ’ cause a malfunction of the gearbox condition monitoring system such as that used to monitor gearbox vibration levels or to detect the presence of foreign or wear-generated contamination. Additionally, other components including gears and bearings may become magnetized due to electromagnetic effects and attract harmful ferrous-based contamination that can lead to wear and premature To mitigate electromagnetic effects, failure. shielding can be used with electrical leads and/or with the electronic component itself. During periodic gearbox inspections and teardowns, procedures often include passing bearings and gears through demagnetizing coils to preclude rebuilding the gearbox with magnetized components. Electromagnetic interference testing methods, such as those prescribed in MlL-STD-462 for military aircraft, will ensure compatibility between all aircraft electrical and electronic components. 6.8 Nuclear, biological, effects
and chemical (NBC)
Reduction in mechanical properties can lead to permanent deformation of the gearbox resulting in degraded gearbox performance or causing oil leakage paths. Qualification tests of sufficient duration should be conducted on the complete
Some aerospace system vehicles may have to operate in a strategic or tactical nuclear warfare environment. Therefore, all system equipment including the gearbox must be optimized to withstand that environment. Potential nuclear weapon effects include the following: blast/gust, thermal, transient radiation, and electromagnetic pulse (see 6.7). In addition, chemical and biological weapons are becoming more widely available and have been used in recent warfare. Thus, the gearbox must be capable of sustaining operation in a total NBC environment. NBC contamination can have direct effects upon the critical properties (e.g., physical, chemical, mechanical, thermal, or electrical) of the In addition, materials or materials employed. procedures used to decontaminate equipment items may also degrade the properties and reduce
gearbox assembly, includingappropriatefittings, to
gearbox/drive system capabilities. To overcome
validate its compliance with specification fire
NBC effects, selection of materials known to be
In the design of a gearbox, materials and fire resistant coatings are selected that offer adequate resistance to fire and which can also maintain appropriate mechanical strength properties. As examples, aluminum and magnesium, which are often used for gearbox housings, have repeatedly demonstrated more than adequate fire resistance and sufficient resistance to thermally degraded mechanical properties.
29
AGMA 9ll-A94
resistant to the expected NBC condition or specific design protection, such as nuclear hardening, must be provided. Material resistance can be established either through testing or by analysis of effects on similar materials or designs. Use DOD-BTU-21 69 for reference and additional information, especially for military applications.
7 Vibration and noise 7.1 Causes of gear vibration Gearbox vibration is the result of a complex interaction among the gears, bearings, shafts, and housings which make up a gearbox assembly. In order to reduce this vibration, the precise mechanisms by which it is generated must be understood. Several mechanisms have been identified which are primary contributors. There are also some secondary contributors which may amplify or reduce vibration levels. Figure’12 is a schematic representation of the way vibration and noise passes from the source at the teeth to the foundation and the air. Note that the path includes the gear bodies, shafts, bearings and housing.
7.1 .I Transmission error One of the most significant sources of gear vibration is the result of the nonuniform transfer of torque between mating gears due to variations in the mating parts, the elastic deformation (bending and contact) of the gear teeth under load, and the deformation of the tooth backup rim and web. The variations can include pitch line runout, tooth spacing variation (both tooth-to-tooth and accumulated pitch errors), and profile variation. Each type of variation will produce a characteristic vibration pattern. For example, pitch line runout will produce a strong once per revolution excitation while spacing variation will produce excitation at mesh frequency (rpm times number of teeth) and harmonics of mesh frequency. The resulting nonuniform transfer of torque produces a dynamic force which in turn excites torsional/lateral/ axial vibratory modes of the gear shafts. The shaft response produces displacements at the shaft support bearing locations. These excite the housing and cause it to vibrate at all mesh frequencies. The dynamic characteristics of the internal components may magnify this excitation. Furthermore, the dynamic characteristics of the housing may also amplify its displacements and resulting noise.
r noise
1
e-’
Airborne structure noise
m
n
Structure-borne noise
Figure 12 - noise and vibration paths 30
AGMA 911-A94
Due to deterioration after many hours of service, many gearboxes will exhibit increases in noise or vibration levels relative to the levels when they were new. This effect is usually the result of changes occurring in tooth profile accuracy and even in tooth spacing accuracy. These changes are the result of surface deterioration due to either wear or localized pitting, and are most common with through hardened gears, but can occur even with case carburized teeth. The condition is aggravated if the operating oil film is marginal, either due to low viscosity or low oil film thickness. 7.1.2 Unbalance All rotating machinery is subject to vibratory excitation due to the dynamic unbalance of the rotating components. Any unbalance will produce a rotating force with a frequency equal to the rotational frequency of the component. This force will be proportional to the square of the rotational speed and will cause a response from all the parts of the gearbox which depends on their dynamic characteristics, i.e., natural frequencies and damping. 7.1.3 Misalignment with connected equipment Many gearboxes have flexible couplings to provide both input and output connections. Misalignment of the connected equipment to these couplings is a possible source of vibratory excitation. In general, if flexible couplings are used which are not of a constant velocity type, misalignment will produce a torsional excitation with a frequency of two times the rotational frequency. The gearbox can respond to this excitation just as to any other vibratory source. 7.1.4 Resonance Although resonance is not a vibratory source, since it represents the coincidence of an excitation frequency with a system natural frequency, it is probably the most common cause of vibration problems. Once the gearbox is designed, the frequency characteristics of vibratory sources in the system are well defined. If any of the gearbox components has a natural frequency which is close to one or more of these excitation frequencies or harmonics of them, vibration or noise can occur. One interesting example of this is the so called “beat” phenomenon where a gearbox will exhibit noise at a frequency equal to the difference between a system characteristic frequency and a nearly equal exciting frequency. Another example
is the case where a turboprop gearbox must avoid natural frequencies at or near the propeller blade passing frequency; i.e., number of blades times propeller rpm. 7.1.5 Entrainment Another source of vibration which is generally of sig nificanceonly on very high speed units or those with very wide face widths and moderate to high speeds is that created when the mixture of air and oil which occupies the space between adjacent teeth is forced out by the entrance of a tooth on the mating gear. In a high speed mesh, the speed of the air-oil mixture as it exits the mesh area is many times the pitchline velocity of the gear set and may, in some cases, reach the speed of sound. At the speed of sound, this effect dominates the overall noise level of the box. 7.2 Consequences of vibration 72.1 Structural issues 7.2.1 .l Fatigue One of the most severe consequences of excessive vibration in a geared system is the possibility of fatigue failure of internal components due to vibration induced stresses. When a critical component has a resonant frequency within the operating range there is a risk that high stressescan occur and some method of reducing the response should be undertaken. This can include redesigning the gear to change its natural frequencies or adding some kind of damping. 7.2.1.2 Fretting Fretting occurs when parts are in intimate contact and are subjected to microscopic motion. Although fretting can occur even if vibration levels are low, the risk of serious fretting problems is greatly increased when the parts are also subjected to high vibration. 7.2.1.3 Fasteners and lockwire A high level of vibration can be a significant contributor to locknut loosening, breakage of lockwire or broken tab washers in gearboxes. 7.2.1.4 Brackets If external brackets are used to provide support for accessories or for gearbox mounting, they can be particularly susceptible to vibration. Since brackets are often designed with simple structural shapes connected together, they can have high stress con-
31
AGMA 91%A94
centrations at the intersections of these shapes. When subjected to vibratory loads, these intersection points can become highly stressed areas.
b) Natural frequency analysis of the gear blanks to determine the mode shapes and frequencies of these components;
7.2.1.5 Lube system components
c) Determination of the dynamic gear loads applied to the components;
Internal lube system components such as nozzles and windage baffles often have low natural frequencies and as such can be susceptible to vibration problems. Particular attention should be paid to attachment points and any bolt flanges. 7.2.2 Noise There is a very close relationship between mechanical vibration and audible noise. If all mechanical vibration were eliminated, noise also would be eliminated. However, it is not necessarily true that reducing the vibration of any single component will reduce noise. On the other hand, selectively increasing the vibration of certain components, if coupled with node relocation, can produce a noise reduction. The noise which is ultimately radiated is composed of two components: airborne and structure borne noise (see figure 12). Airborne noise is transmitted directly from the source. Structurebome noise is due to the transmission of the vibratory energy through the internal structure of the gearbox and into the external supporting structure. The external structure provides a path for noise since it may be excited by the gearbox at its attachment points and produce or amplify noise. 7.3 Design 7.3.1 Finite element analysis The ideal time for minimizing vibration is in the preliminary design stage. The ability to analyze a given gearbox and modify its design, based solely on this analysis, in order to minimize its operating vibration level requires the use of several finite element modeling techniques. These analyses define the excitation due to the gears, the response of the shaft support system to these excitations, the manner in which these shaft responses are transferred to the housing through their bearings, and the response to these various stimuli. In general, the approach involves the following analyses:
a) Modeling the gearteeth for local dynamicflexibilii and kinematic loading;
32
d) A detailed finite element model of the static gearbox structure; e) An analysis of the modes of the entire system. 7.39 Strain energy considerations In recent years, a trend to dynamic optimization by strain energy techniques has evolved. This has mainly been for the alteration of an undesirable natural frequency. To understand the technique, consider that in general each natural mode of the structure contributes to vibration in proportion to its amplification factor. Consequently, each mode whose frequency is in the vicinity of the forcing frequency would be a major contributor to the overall dynamic response. In the modal method, which operates principally on the amplification factor, the natural frequency immediately above the exciting frequency is usually increased. One could also reduce the natural frequency immediately below the exciting frequency if it is possible structurally. A finite element analysis is first employed in the modal method to yield a dynamic solution. The mode shapes are obtained, then the modal strain energy distribution throughout the structure is found for any given mode shape whose natural frequency is to be modified. The strain energies for all structural elements are obtained and then tabulated from the highest to lowest. The structural elements with the highest strain would be the best candidates for modification of the natural frequency. For example, in the case of increasing the lowest mode, the elements with the highest strain density, when deformed in this mode, would be the best candidates for modification to obtain a maximum frequency shii for a minimum material addition (weight) penalty. 7.3.3 Design parameters Inordertoprovidesomefeelforthemannerinwhich various transmission design parameters affect vibration and noise level, some of the more significant ones will be addressed in the next subclauses.
When utilizingthis information,the designer should beaware that they represent trends only. Due to the
AGMA 91%A94
complex interactions which exist in a gearbox, the results of similar investigations utilizing significantly different test boxes may vary substantially. 7.3.3.1
Tooth combinations
The selection of tooth combinations which avoid potential resonances is sometimes possible. By varying the selected tooth numbers, the meshing frequency can be raised or lowered. Obviously, this must be done within the limitations imposed by stress and scuffing (scoring) requirements. 7.3.3.2
Contact ratio
The properties inherent in certain types of gears can have a large effect on gearbox vibration levels. Accuracy aside, contact ratio (simply stated, contact ratio is the average number of teeth in contact during each mesh cycle) is one of the most important parameters which affects gear tooth excitation and thus noise and vibration level. Virtually all other gear tooth parameters which affect vibration level do so largely by virtue of their effect on contact ratio. For example, decreasing a spur gear’s pressure angle generally decreases noise level, however the same effect can be achieved by extending the addenda of the higher pressure angle gears slightly to achieve the same contact ratio as the lower pressure angle gear. Unfortunately, due to tooth geometry limitations, this approach can be used only for relatively small changes in pressure angle. As noted earlier, the primary exciting force within the gear mesh is due to the non-uniform transfer of load between successive pairs of meshing teeth. Increasing the number of tooth pairs in contact reduces the total load carried by any single pair of teeth, thus the dynamic forces generated at engagement and disengagement are reduced. In addition, as the number of tooth pairs in contact increases, individual tooth errors tend to average out, again reducing the dynamic loads generated. One way to improve contact ratio and thus reduce noise is to use helical rather than spur gears or spiral bevel rather than straight or Zero1bevel gears. Generally, when considering vibration and noise levels, the designer should strive to achieve the highest contact ratio possible within the constraints of the system being designed. Since vibration or
noise is seldom the single driving force in the design of a transmission system, other considerations will dictate the basic type of gearing to be used. Given that the box is to be of the parallel shaft type, the designer may choose from simple spur gears (contact ratio 1.25 - 1.70), single or double helical gears (total contact ratio* 1.8 - 5.0, or greater in some cases), high profile contact ratio (HCR) spur gears (contact ratio greater than 2), or any of a variety of special purpose geartypes. The choice of which specific type of gearing to be used depends on many considerations, only one of which is its vibratory characteristic. 7.3.3.3 Tooth shape modification In many cases, it is desirable to modify the gear tooth profile shapes so that they are no longer true involute curves, especially for high load, high speed drives. This is accomplished by relieving the tip andor the flank of the profile. This is done to avoid tooth interference on engagement and disengagement during rotation. This interference is due to deflections, both within the gear teeth themselves and in the housing and shafts. If these interferences are not relieved, the load capacity of the gears can be seriously impaired. In addition, since these conditions generate high dynamic loads, vibration excitation and noise is also increased. In addition to modifying the tooth profile shape, it is also often desirable to modii the tooth shape in the axial direction. This can take the form of crowning or lead correction. Lead or helii corrections compensate for deflections of the teeth and shafts and also for the housingsand supports. If pinions have a high length to diameter ratio (above l/i), and are subjected to heavy loads, they will often have enough “windup” to cause heavy end loading if compensation is not provided. This heavy concentration of load at the ends will contribute to vibration and noise. Crowning may also be used to relieve end load. It must be emphasized that any tooth shape modification is optimum for only a single load level. Vibration levels will increase both above and below this load level. If the modification is not very severe, the increase at lower loads may not even be perceptible but heavily modified gears will be noisy at offpeak
.* Total contact ratio is equal to the transverse contact ratio plus the face contact ratio.
33
AGMA Qll-A94
loads. For this reason, the actual modification applied will often be a compromise. 7.3.3.4 Gear accuracy It is usually true that a more accurate gear also has less vibration, assuming appropriate tooth modifications. If the gear mesh is exciting a resonance somewhere in the system, nothing short of a uperfect” gear set will substantially reduce the vibration and noise. In most other cases however, improving accuracy will reduce the vibratory levels. This is especially true when relatively low quality gears are being used but much less so for higher quality gears. Among the most important characteristics that affect noise and vibrations are involute hollowness, lead variation, and control of profile modification. 7.3.3.5 Design of shafts and location of bearings The design of the shafts which support the gears and the location of the bearings which support the shafts are a critical concern for two quite different reasons. The first is the need to provide the gears with a support system which maintains their relative alignment. A misaligned gear system is subject to higher unit tooth loads which not only degrade load capacity but also increase the overall noise and vibration level of the box. In many cases, especially for high speed, lightweight gears, the support shafts are frequently sized based on deflection and not solely strength restrictions. The location of the bearing supports, as noted earlier, should be chosen so that they coincide as nearly as possible to node points in the shaft mode shape. This will ensure minimum vibration transmission to the housing and thus minimum noise.
reduce the propagated levels. Often, these ribs can be spaced so that they also contribute to improved cooling. The second approach to housing design is more quantitative. Utilizing the results of the finite element analysis, the strain energy density of the entire housing can be evaluated. With this information in hand, thoseareas whichare highest in strain energy density, and thus the most likely candidate areas for vibration reduction, can be identified and modified. 7.3.3.7 Bearings Vibration or noise generated by the bearings is generally overshadowed by that due to the gears and their interaction with the housing. Vibration created by rolling element bearings can sometimes be reduced by control of manufacture, through tight tolerance and surface finish. This can lead to an increase in bearing cost. Another approach to reducing vibration may be to select the quietest bearings from a production lot. Many bearing manufacturers utilize a sound checkas one of their quality control devices. The bearing to be checked is mounted in a standard fixture and run at a standard speed. its noise level is checked and those which exceed a given value are rejected. 7.3.3.8 Tooth stiffness Several factors affect tooth stiffness, such as pressureangle, andtheratiooftooththicknessto height. In general, the more flexibility the tooth possesses the lower the dynamic loading will be, and thus vibratory excitation will be minimized. Changing these parameters in the directions required to improve flexibility may compromise tooth strength. 7.3.3.9 Backlash
First, the housing design can be evaluated qualitatively. The design should be reviewed to ensure that there is a rigid load path between each bearing location and the housing mounting points to reduce deflections. Large flat or gently curved areas on the housing surface should be avoided since these
Sufficient backlash should be provided to avoid tight mesh (drive and coast side contact simultaneously) under all load and temperature conditions. This backlash should be obtained by tooth thinning and not by spreading the center distance since the former method does not affect contact ratio while the latter does. Under conditions of reversing loads or during periods of zero load operation, excessive backlash will result in gearbox noise. The loss in tooth bending fatigue strength due to tooth thinning must also be considered; however, for most gear
areas tend to vibrate freely (like a drum) when
systems the actual reduction in strength is very
excited. Ribs can be used to stiffen these areas to
small.
7.3.3.6 Design of housing There are two major approaches to housing design.
34
AGMA 911-A94
For very high speed gears or wide face gears with moderate to high speed, an increase in backlash, above standard values, may be required to reduce the velocity of the air-oil mixture which is forced from the tooth space during meshing. Small increases in backlash can result in measurable improvements when the overall noise orvibration level is dominated by this phenomena. Spur gears are more susceptible to this phenomena because of the large trapped volume which exists during meshing. 7.3.3.10 Root clearance The effect of root clearance is similar to that of backlash in that a sufficient amount must be provided to avoid any tip interference with the mating gear. Additional root clearance (increased tooth whole depth) above this value will have the following effects: - Tooth bending stress will be increased since the beam length will be increased; - The smaller fillet radius in the root will also contribute to increased bending stress;
7.3.3.12 Gear material For steel gearing, the choice of a specific material has little effect on thevibration level. If a non-metallic material can be utilized, however, it can have a significant effect. This is due to two factors. First, the increased compliance of these materials decreases the magnitude of the dynamic loads and second, the materials themselves are usually good dissipative dampers. These factors combine to make non-metallic materials a good choice when the operating environment (temperature, load, speed, lubricant compatibility, etc.) permits. A secondary benefti of this approach is that such gears when run in combination with a steel mate may be run with minimal or no lube. 7.3.3.13 Surface finish The surface finish on the flanks of gear teeth has only a small effect on vibration and noise level. 7.3.3.14 Relative influence of factors Table 12 is a summary of the design approaches discussed above showing the relative influence each characteristic has on vibration and noise. 7.4 Analyzing vibration problems
-There will be a reduction in noise caused by the exit velocity of the air/oil mixture. This is due to the reduction in the exit velocity provided by the increased clearance. In most cases, exit velocities are not high enough to be significant; -There will be an increase in tooth flexibility and thus an improvement in load sharing and error tolerance.
7.4.1 The excitation mechanism The initial step in analyzing vibration problems is usually to define the excitation at each gear mesh. The factors described in table 12 can all contribute to this excitation. Table 12 - Potential influence of design features on noise and vibration Design feature
7.3.3.11 Gear mesh phasing Sound transmission is accomplished through the propagation of waves which can be considered sinusoidal. The addition of two sine waves with identical frequencies which are separated in phase would produce a third wave of the same frequency but with a third amplitude. While gear noise may not be exactly sinusoidal it can be phased. Studies of such phasing have proved that a significant reduction in noise level can thus be achieved in applications having multiple load paths. Examples of methods to use phasing to reduce noise and vibration are: tooth selection in multiple path systems, unequal spacing of planet pinions, and change of double helical gear intersection point.
Accuracy Contact ratio Load intensity Phasing Tooth profile modification Alignment Bearing location Bearing type Housing design Tooth numbers Tooth stiffness Backlash Material Root clearance Surface finish
Influence High High High High High Medium Medium Medium Medium Medium Medium Low Low Low Low
AGMA 91%A94
A gear system transmitting power is also susceptible to torsional vibration. The inertia may be concentrated as in the body of a gear or distributed as in the shafting. Similarly, the elasticity may be concentrated as in a coupling, or it may be distributed with the inertia in the shaft sections. The excitation may come from externally applied pulsating torques or from a fluctuating resistance to the steady rotation. However, in a geared system there is also an excitation due to displacement which comes from the imperfect transfer of motion (as described in 7.1 .l) between the meshing gears. Having defined the gear tooth excitations, a modified Holzer analysis can be used to calculate the dynamic tooth forces [9]. 7.4.2 Response characteristics 7.4.2.1 Shaft response Once the dynamic forces at the gear mesh have been calculated, they can be applied to an analytical model of the shaft system. The results are dynamic forces at the shaft support (bearing) locations and vibration mode shapes of the shafts. The objective of the shaft analysis is to reduce the shaft response in general and to change the shaft geometry such that the bearings are located close to node points for the critical modes. 7.4.2.2 Gear and housing response The final step in the analysis is the evaluation of the response of the gears and housings to the excitations transmitted to them. This task is usually accomplished through the use of a finite element program. The responses are obtained by applying the appropriate forcing functions at the tooth contact points in the case of the gears or at the bearing locations within the housing. In the development of the model, the parts are supported at their normal mount points by appropriate constraints [lo]. 7.4.3 Test methods Many different methods of testing are used to determine vibratory characteristics. Each method is designed to answer certain kinds of questions, but no one test will resolve all vibration or noise issues.
loading can be applied. This method offers the advantage that the parts are subjected to the operating dynamic conditions, so that results will closely resemble field experience. 7.4.3.1.2 Siren This is usually a non-running test in which the part to be tested is excited by variable frequency air pulses. A stream of pressurized air is passed through a rotating serrated disk whose speed can be varied over the frequency range of interest. The air is directed at a location which would excite the mode shapes of interest. Since several diierent disks can be used with differing hole patterns, frequencies up to about 20 kHz can be investigated. The response can be determined either by using a camera if there is sufficient motion or with strategically located strain gages. Since this method applies a low dynamic force to only one specific component, it is not usually practical to simulate the response of a complete gearbox by this technique. 7.4.3.1.3 Impact This is also a non-rotating test in which the component to be tested is supported in a manner which will not affect its frequency response, and then is struck with an impact device, typically a hammer with a dynamic force gage, selected for the frequency range of interest. The response is measured with accelerometers or strain gages placed at critical locations. By impacting the part in various locations and monitoring the responses, it is possible to determine natural frequencies and mode shapes. However, since it is a static test, any change in frequencies or damping characteristics due to the effect of rotation is not included. 7.4.3.1.4 Shaker This type of test is similar to the impact test except that an electrodynamic shaker head is mounted so as to impart a force to the component being tested. The force can be varied in frequency so that the response can be determined at critical frequencies, either with accelerometers or other transducers. Since it is a non-rotating test, the same cautions apply as with impact testing.
7.4.3.1 Excitation methods
7.4.3.2 Response measurement
7.4.3.1 .I Running tests
74.3.2.1 Strain gages
If the parts to be tested can be operated under realistic conditions, simulated or actual service
When one of the parameters to be studied is the stress in the part during vibratory response, strain
36
AGMA 911-A94
gages can be applied to locations which are expected to show high stresses. The gages are used to determine both the frequency response and the operating stresses. For the results to be useful, the excitation method must simulate operating conditions. 7.4.3.2.2
Accelerometers
Since the application of strain gages can be diiicult and time consuming, accelerometers are often used as the transducers. Similar results with respect to frequency response and mode shape can be obtained except with no information relative to stress. The availabilii of small, lightweight accelerometers has made this technique very popular. 7.4.3.2.3 Powder
pattern
This technique for determining modes and frequencies involves coating the part with a special powder and exciting it with a variable frequency shaker. By observing the patterns developed in the powder at the part’s natural frequencies, the modes for those frequencies can be determined. It is also a nonrotating test, so effects of rotation are not included. The part is normally excited during the observation by a shaker, since the air siren would interfere with the patterns. 7.4.3.2.4 Holography This is a laser based optical technique of observing and photographing mode shapes. The components are excited, usually by a shaker, at the selected frequency. The holographic camera records a hologram of the part which is vibrating during the exposure. The resulting image is covered with light and dark fringes which are contour lines related to the vibration amplitude. From this fringe pattern, the mode shapes can be determined and, if desired, the pattern can be quantitatively evaluated to determine vibration ampliiude at any point. The approach is most useful for non-rotational determination of modes, since obtaining a fringe pattern on rotating parts requires an elaborate setup. Table 13 lists some common combinations of these techniques and shows the characteristics of each. 7.5 Vibration/Noise
reduction
techniques
In cases where hardware is in existence before a noise or vibration problem has been recognized and
addressed, corrective measures must center around modification to the existing hardware. Several of the more effective methods will be discussed. Table 13 - Vibration
testing
Test objectives Excitation: Measurement
Running: Strain gage Accelerometer Siren: Strain gage Holography Impact: Accelerometer Shaker: Holography Powder Strain gage 7.5.1
g
yes
X x
X x
no no
X X
X X
no
X
X
no Yes no
X
X x X
X
X x
x
X
X x X
Isolation
Isolation means that excitation forces are prevented from being transmitted from one part of the system to another. This interruption yields a reduction in the transmitted forces. It is most common to isolate either the transmission from its supporting structure or the gear rim from its blank and hub. The former approach is generally easier to accomplish and is therefore the most common. 7.5.1 .l
Gearbox
mounting
isolation
Vibration isolators are often used to control the transmission of high frequency energy across the mounting points to the supporting structure. This method involves the use of elastomeric mounts between the transmission mounting pointsand its supporting structure. Reductions as high as 10 dB can be obtained through the use of rubber mounts having a static spring rate equal to that of the supporting structure at the gear box attachment point. 7.5.1.2
Gear rim isolation
A second method of reducing vibration and noise is to ‘s olate the gear rim from its hub with an elastomeric insert. The elastomer will reduce blank excitation and, consequently, shaft and housing exci-
37
AGMA 911-A94
tation are also reduced. One of the difficulties with this approach is the limitation on torque transmission imposed by the strength of the elastomer. The effect of the lubricant on the elastomer and the accuracy of the position of the gear with respect to its mate may also limit its use to non-critical, low load situations. In those systems where load capacity is secondary and vibration or noise is primary, this method has been shown to work quite well. 7.5.2
Dissipation
7.5.2.1 Externally
applied
damping
material
Application of damping material to the surface of a structure by spraying, gluing, plasmaflamecoating, etc., may be effective. Since most of the straining action (and thus energy conversion into heat) will be confined to the layers closest to the structure, thin layers of damping material will be more cost and weight effective than thick layers. Objections to the use of damping materials applied to the exterior surface of transmission housings . include greater cost, added weight, concealment of cracks, and heat retention. Each of these must be traded off against the potential benefits. There are many methods by which dissipative damping may be achieved. The simplest of these is to fill a hollow gear shaft with an elastomeric material, or to provide constrained layer damping. Dissipative damping is effective when constrained layer damping is applied to gears or housings with large flat surface areas which tend to “ring” when excited. 7.5.2.2 Damping _ - rings Rotating parts such as bevel or spur gears have many natural modes of vibration which can be excited within the operating range. It is often impractical to redesign the part to avoid all of these frequencies, so some means of reducing the response is the next best solution. The use of damping rings is quite common in this situation. These can take the form of special snap rings installed in grooves in the gear rim, or more complex rings designed specifically for the purpose. Figure 13 shows an example of a typical damping ring for a spiral bevel gear rim. The location of the grooves and the shape of the ring are determined by the mode shapes of concern and by the need for the
38
frictional damping to occur at the interface of the rings and grooves. Care must be taken to avoid having the rings become a wearing part and thereby introducing an additional mode of failure.
Figure 13 - Typical damping 7.5.3
ring
Screens or barriers
A well constructed screen can shield a high frequency source and result in a significant noise level reduction. However, if the gear drive is operating at avery low speed, it will produce IowfrequWcy noise and the gap around the screen can act as a new source and radiate sound into the space with little attenuation. Some noise will still enter the working area via reflection from the adjacent walls as well as by transmission through the-screen itself. In this case, the use of absorbent materials, in conjunction with a screen, can prove quite effective. 7.5.4
Absorbent
materials
Sound absorbing materials serve the purpose of minimizing reflection of sound waves. They are not very effective at blocking the path of sound transmission. For this reason, they are usually used to alterthe sound characteristics of an area, and in this manner, reduce the sound pressure level within that area. This type of material is not normally used alone as a barrier between a source and a receiver, since the sound pressure level would be reduced very little at the receiver. Some factors affecting the absorption characteristics of a material are mass, surface condition, pore size and structure, flow resistance, thickness and the frequency of impinging sound. Typical materials used are glass fiber and certain types of polyurethane foam. 7.5.5 Enclosures Reduction of airborne noise by the use of enclosurescan be effective, but the noiseattenuation that can be achieved is dependent upon the complete-
AGMA 911-A94
from the stress index prediction. The reasons for disagreement between actual stresses and stress index numbers can be caused by any number of factors including the following:
ness of the enclosure. The noise reduction limitation in the speech frequency range with typical acoustical enclosures and seals is about 25 dB, with up to 35 dB obtainable through use of improved seal configurations. Further reductions in noise level, up to 50 to 60 dB, can be achieved with fume-tight enclosures, such as those employed in some commercial helicopters and in some commercial transport aircraft engine installations operating today.
- Size effects; - Fine pitch vs. coarse; - Blank configuration; - Rim; -Web; - support;
Practical enclosures are limited in noise attenuation by unavoidable sound leaks in seams and access doors. Not only do these enclosures impose considerable weight and maintainability penalties, but they do not reduce the harmful effect of the accompanying vibrations which contribute to material fatigue and fretting at joints.
-
A summary chart showing the relative effectiveness
of the various methods discussed given in Table 14.
in this clause is
Table 14 - Noise and vibration reduction techniques I Relative effectiveness
Mounting isolation Rim isolation External damping Internal damping rings Screens Absorbent material Enclosure
Airborne noise
Local vibration
Medium Medium Medium Low Low Low MedIHigh
MedIHigh MedlHigh High High Low Low Low
8.1 Introduction considerations
for fatigue
The conventional AGMA method of bending stress calculation is in reality a stress index method of design. Production designs based on this method have countless hours of successful field operation. The allowables are based on this successful field experience and successful/unsuccessful bench test experience. When the stresses are determined
experimentally
by strain gaging gear teeth, it is
found that the actual stresses may be quite different
Thus the designer faces the dilemma of whether or not to use the stress index method or to conduct a more complex analysis, such as finite element analysis. This is not to say that the classical method is inaccurate for it can be shown that the final application factor by the stress index method and FEA are in agreement if the proper values for the factors have been chosen.
I
8 Load capacity 8.1.1 Analytical bending
Dynamic effectslresonance; Temperature effects; Loading conditions/prelubing; Deflections of teeth, shaft, housings.
The conventional AGMA method permits calculation of the maximum tooth tensile stress. The minimum stress is assumed to be zero. Thus the vibratory and steady stress is taken to be l/2 of the maximum tensile stress. The true minimum bending stress in the gear is usually negative and occurs prior to the initial point of contact for a driving gear and after the final point of contact for a driven gear (see figure 14). Afinite elementanalysiscan predict the maximum and minimum stress by conducting a series of static solutions of stress as the pinion and gear are incrementally “rolled” through mesh. The stress allowables for the finite element method are generally taken to be the material allowables found from material testing or from actual gear teeth test results. A design approach that has been employed is to conduct the design using the classical AGMA stress index method and then to conduct an optional finite element analysis to refine the design. A digital computer analysis using AGMA equations can quickly and efficiently compare many preliminary designs
whereas the finite element method is very tedious and time consuming.
On the other hand, the FEA
39 39 39 39
AGMA Qll-A94
Load
A
A
Time Maximum stress, thick rim (AGMA stress index) Load
Load
Time
Minimum and maximum stresses, thin rim (FEA method) Figure 14 - Different methods for determining tooth root stress can more readily account for variables such as suppottstiSfnessand temperature effects. Accounting for these factors permits the load distribution to be determined. In aerospace gearing where the gear rims are designed to minimum thickness, the gear rim and web can have a profound effect on the bending stress. Thin rims make the calculated stress number too low, giving an optimistic estimate of life. As the thickness of the gear rim is reduced, there is a point where the stress in the gear rim itself is higher than the tooth root bending stress. This is generally not a good design condition because cracks may propagate through the gear rim instead of through the tooth. It is of course not desirable to have any cracks which propagate, but assuming a crack were to be present, it is desirable to control the direction of propagation. Cracks through the gear rim should be avoided because they can cause sudden and catastrophic loss of the mesh without warning since chips are usually not generated by this type of fracture. Hence the stress in the root of the tooth should be lower than the tooth bending stress which is usually maximum at a position near the TIF. Stresses in the gear rim
40
and web can be calculated by FEA. The user should satisfy himself that the gear blank construction is representative of the basic theory embodied. Figures 14 and 15 illustrate the above discussion.
propagation through tooth
1
Maximum tooth rim stress
9 Direction of propagation through rim Figure 15 - Directions of crack propagation in gear teeth
8.12 Design criteria Prior to beginning the gear analysis several design parameters must be established. In aerospace gearing, the design is usually based on three surface bending fatigue, considerations: compression (Hertz stress), and scuffing (scoring) resistance. Each of these conditions must be
AGMA Wl-A94
checked independently for adequacy of design. For aerospace designs which incorporate condition monitoring, bending tooth breakage probably has the most severe consequence of the three failure modes whereas pitting and scuffing are durability type failures. Durability type failures can cause gearbox removals but are not catastrophic unless the gearbox is permitted to operate in that condition for a sufficient time period in which the surface breakdown degenerates to the point where the mesh is lost. Pitting and scuffing are a phenomenon of the mesh while bending is individual to the pinion and gear, therefore both members must be analyzed for bending. Optimization of scuffing and a bending stress balance or life balance between pinion and gear can be achieved by varying addendums and dedendums and/or tooth thicknesses of both members. In aerospace designs, it is desirable to achieve a balanced tooth design. That is to say that the life in bending, contact, and scuffing resistance should be optimized. This is not always possible for a number of reasons, but the gear designer generally has control of basic parameters such as the number of teeth, diametral pitch, and center distance. In general, the gearset is designed for Hertz stress (for single external mesh) as follows: 315ooP (V&+ 1)3 C2F
...(9)
= A- nP “G 6.923 x lo8 P (mG+ 1)3
62F = ’
.*..(9M)
nP mG
From equation 3.1 of ANSVAGMA 1012-F90, C=
. ..(lO)
and if F is set equal to d, and substituted in 9, the cubic equation can be solved ford, (and F), where c
is center distance, in (mm);
dp
is operating pitch diameter of the pinion, in (mm);
F
K
is face width, in (mm); is contact load factor for pitting resistance, lb/in* (N/mn?);
P
is transmitted power, hp (kVV);
w
is gear ratio;
np is pinion speed, rpm. NOTE:K= 500400 istypicalforcontinuousoperation of carburizedaerospacegearing. The teeth and diametral pitch can be selected for bending. The gearset scuffing resistance is then analyzed. 8.1.3 Design life considerations Of primary importance in the preliminary design are reliability requirements. The life can be calculated using Miner’s cumulative damage theory (see Appendix B of ANSVAGMA 2001-B88 for explanation of Miner’s Rule). The gear designer must have a spectrum of estimated operating power, rpm, and percent time. This usage spectrum should be established by analysis of the operating environment and intended use of the gearbox being designed. Even in the case of relatively slow speed drives, the cycle build up is very rapid. Unless operation at the highest powers is for very limited durations, finite life design becomes impractical. On the other hand, if the high power condition is truly transient in nature and the rpm is relatively low, finite life design can achieve reduced size and weight. To conduct a finite life design, material stress vs. cycle curves must also be known (see ANWAGMA 2001-B88 for allowable stress and life factors). If it is desired to have unlimited life, the gearing can be designed for limit load or the maximum load expected in service. 8.2 Spur, helical and bevel gear tooth breakage and surface durability The bending strength of gear teeth is a fatigue phenomenon related to cracking at the tooth root fillet. See ANSVAGMA 110.04-1980 for a more complete discussion. The basic theory assumes that the tooth is a cantilever beam fixed at its base (the gear rim) with a reduction in tensile stress caused by the compression of the radial component of the transmitted load. A stress concentration factor is applied because the tooth is subjected to alternate cycles of maximum to zero stress. Load sharing between adjacent teeth in contact is calculated. The pitting of gear teeth is a fatigue phenomenon caused by repeated application of surface com-
41
AGMA 91%A94
pressive stresses as the teeth roll through mesh. Initial pitting and destructive pitting are illustrated and discussed in ANSVAGMA 110.04.
SC = s
J
w* G C,
cs dTe
cnl Cf I
. ..(12)
where 8.2.1 Fundamental bending stress formula (Refer to ANSVAGMA 2001-888 for spur or helical gears or ANSVAGMA 2008-A86 for bevel gears) 5 --v wt ‘-K,,K
Ka
pd
Ks
F
Km KB
J
S_ dp
ca cv
...(n)
cs dP
s _ Wt Kca 1 5 t ?TlF KVKX
Km KB J
. ..(ll M) Fe
where
KB KV
K,
J pd
m F
is calculated tensile bending stress number at the root of the tooth, lb/in* (NImrr?); is tangential tooth load, lb (N); is external application factor for bending strength. This factor takes into account the effect of any externally applied load in excess of the nominally applied torque load; is the rim thickness factor for spur and helical gears (use 1 .Ofor bevel gears); is dynamic factor for bending strength. This factor takes into account the effect of gear tooth quality, as related to speed and load; is size factor for bending strength. This factor takes into account nonuniformity of material properties; is load distribution factor for bending strength. This factor modifies the rating equation to reflect the non-uniform distribution of load across the face of the tooth; is tooth lengthwise curvature factor for bevel gear bending strength (use 1 .O for spur or helical gears); is tooth geometry factor for bending strength; is diametral pitch, in-l; is module, (mm); is face width of gear for which bending stress is desired, in (mm).
8.2.2 Fundamental contact stress formula Refer to ANSVAGMA 2001-888 for spur or helical gears or ANSVAGMA 2008-A86 for bevel gears.
42
% Cf
is contact stress number, lb/in* (N/mm2>; is elastic coefficient, W (m ); is application factor for pitting resistance; is dynamic factor for pitting resistance; is size factor for pitting resistance; is operating pitch diameter of pinion, in (mm); is effective or net face width of narrowest member, in (mm); is load distribution factor for pitting resistance; is surface condition factor for pitting resistance;
I
is geometry factor for pitting resistance. 8.2.3 Dynamic factor A dynamic factor is applied to the tooth bending and contact stress to account for increased dynamic tooth load. lnertias and spring rates of the transmission system influence the dynamic tooth load as does operating speed compared to resonance speed. Gear tooth tolerances have a large effect on dynamic load. As the gear tooth rolls through mesh, teeth are engaging and disengaging as a function of the tooth geometry. For example in a conventional mesh, there are alternate cycles where the load is carried by varying numbers of teeth. In the worst case, since the spring rate of the mesh is considerably different when one tooth is meshing and when two teeth are in mesh, the tangential deflections are different at different times during the mesh cycle. This change of mesh stiffness with time causes the mesh point tangential deflection to vary with time at the tooth mesh frequency. The resultant change in tangential deflection caused by a change in velocity from the theoretical constant velocity is sometimes referred to as transmission error or TE. Transmission error induces a vibratory forcing function at the mesh point and is the root cause of dynamic tooth load and noise. Profile modifications on gear teeth are required to accommodate tooth mesh deflections and other
AGMA 911-A94
variables. Modifications can also have an effect on tooth spring rate. The modification specified is optimum for only one load condition, usually some high load, high speed operating condition. When operated at other load conditions, the gearset may produce higher dynamic loads. Excessive profile modification can actually reduce mesh contact ratio and introduce error in action. A complete discussion and calculation of transmission error and resultant dynamic load is beyond the scope of this guide. A simplified dynamic factor can be obtained by referring to ANSVAGMA 2001-S88 for spur or helical gears or ANSVAGMA 2003-R% for bevel gears. 8.2.4 Siie factor Size effect factor has been established to account for detrimental effects as the volume of material in the gear set increases. This can be thought of as a chain analogy. The more links in the chain (units of stressed volume), the higher the probability of a defect in one of the links. For aerospace spur and helical gears the gear size has not been established to have a detrimental effect and a factor of 1.O is used. For bevel gears a size factor is used and can be determined by referring to ANSVAGMA 2006-A86. 8.2.5 Load distribution factor The load distribution factor accounts for non-uniformity of load across the lines of contact on the teeth. The deviation from the theoretical uniform load is caused by operating deflections and dimensional allowances. Deflections along the teeth are caused by the following: - elastic deflections of the gear elements; - shaft deflections; - bearing deflections; - housing deflections; - foundations which support the gear elements; -thermal expansion from non-uniformity of temperatures; - differential thermal expansion from different materials; - centrifugal effects.
Dimensional allowances that affect load distribution are as follows: - accumulation of tolerances; - alignment of the axes of rotation; - bearing clearances; - profile accuracy; - lead; - crowning. For a particular application, each of the above influences should be evaluated as to its effect on load distribution. Modem finite element methods can be used if the pinion and gear are modeled as separate parts and the tooth load induced by torque applied to the shafts through the use of gap elements or with three dimensional contact modeling. The finite element technique is suitable for calculating load distribution because all of the factors which influence deflection and manufacturing deviations can be evaluated. 89.8 Geometry factor, J The geometry factor is used to calculate gear tooth bending stress. The geometry factor takes into account bending of the tooth as a cantilever beam as well as the compression caused by the radial component of the normal tooth load. The maximum stress occurs where the load is at the highest point of single tooth contact for a conventional low contact ratio gear tooth. For high contact ratio gears a load sharing factor is applied to account for distribution of the transmitted load among the teeth. Internal gears are often neglected in texts but are important in aerospace design because of heavy reliance on planetaries which have an internal ring gear. Annex A shows a procedure for calculation of the geometry factor for an internal gear mesh. The procedure is suitable for calculation of either internal or external gears or mates. The equations are generalized by the use of “signed integers” which are +l for external gears and -1 for internal gears. Helical gear geometry factors are based on a mean normal tooth section and also account for load sharing. After the mean normal section is established, the geometry factor iscalculated in the same manner as for a spur gear. See Annex A for spur gear geometry factors. Further data on the calculation of spur and helical gear geometry
43
AGMA 911-A94
factors is shown in AGMA 908-B89. For bevel gear geometry factors, see ANSVAGMA 2003-A86. 8.2.7 Allowable bending stress The allowable bending stress is calculated as an allowable stress index number and represents the allowable stress when the stresses are calculated by the procedure outlined above. The allowable stress number is a function of the material and heat treatment used as well as the desired life, temperature of operation, and reliability desired. The relationship is given by: Swt
I-
Ct KL
...(13)
KTKR
where sWt is working bending stress number, lb/in* (N/mm*); S
at
is allowable bending stress number, lb/in* (N/mm*);
KL
is life factor for bending strength;
K,
is temperature factor for bending strength;
KR
is reliability factor for bending strength.
The reliability factor accounts for the normal statistical variations found when materials fracture. From the analysis of probability, for a normal distribution, the reliability constant is given as: -Rel = 1-nv where Rel
n V
...(i 4)
is reliability constant; is number of standard deviations; is coefficient of variation = standard deviation/mean.
It has been found experimentally that for steel, the coefficient of variation is approximately 10% (v = 0.1). This number can be used for spur and helical gears but the variation in spiral bevel gears has been found to be higher because of problems encountered in shimming or other dimensional variables which influence gear tooth patterns. A standard deviation of 0.156 has been calculated for bevel gears using data from a large number of test
16. For highly reliable aerospace design, a reliability of 3 standard deviations has been used in the past (or 3o). This results in a reliability of 0.99875 and a reliability constant of 0.7. Thus the working 3o allowable is found by multiplying the mean stress by 0.7. From this discussion it is seen that there can be different allowable design stresses for the same design and that each stress will have a corresponding reliability associated with it. It is up to the designer to establish reliability goals before the design begins so that allowable stress can be determined. The allowable stress numbers shown in ANSVAGMA 2001-B88 are based on a reliability of 1 failure in 100 at 1Oscycles. This corresponds to n = 2.326 (number of standard deviations corresponding to 99% reliability). Thus for a spur or helical gear with v = 0.1, the reliability constant for 99% reliability is 0.7674 and for a bevel gear with v = 0.156, the reliability constant is 0.6371. To determine the reliability factor for multiplication of the stress index:
KR
=-
R99
. ..(15)
Rdesired
where R99
is reliability constant for 99% (0.7674 for spur or helical gears and 0.6371 for a bevel gear);
Rdesired = 1-nv = reliability constant for desired reliability. The allowable bending stress numbers and factors are shown and discussed in ANSVAGMA 2001-888 for spur and helical gears and in ANSI/ AGMA 2003-A86 for bevel gears. Reliability (or probability) is related to the number of standard deviations by the term: n & (in (...16) Rel = s Unfortunately, there is no closed form solution to this integral. However, it may be solved by a
and field bevel gearfractures. A plot of reliabilityvs.
numerical approach. For a numerical approach to
number of standard deviations is depicted in figure
calculate reliability, knowing the numberof standard
44
AGMA 91 l-A94
deviations, n:
Rer= 1 (...17)
8.3 Spur, helical, and bevel gear scuffing (scoring) - flash temperature index
where
t
bending stresses. The allowable contact stress numbers and factors are shown and discussed in ANSVAGMA 2001-B88 for spur and helical gears and in ANSVAGMA 200%A86 for bevel gears.
1 = l+
8.3.1 Fundamental flash temperature formula
0.2316419n
Cl = 0.319381530
The fundamental scuffing (scoring) formula is derived by the method proposed by Blok which calculates a temperature at the gear mesh point as the gear bulktemperature added to the temperature rise in the mesh. The temperature rise in the mesh is referred to as the flash temperature index. Appendix A of ANSUAGMA 200%B88 gives a detailed discussion of the flash temperature index method for spur or helical gears, but there is no generally accepted AGMA method for bevel gears.
c:! = -0.356563782
1.781477937
c3 =
c4 =- 1.821255978 c5
1.330274429
=
n = number of standard deviations (horizontal axis of figure 16)
Rel=
reliability (vertical axis of figure 16)
By the above method, Relwill be accurate to 4 decimals. 8.2.8 Allowable contact stress The allowable contact stress is a function of parameters similar to those used for allowable bending stress number and in addition includes a hardness ratio factor. The relationship is given by: SWC
I
sac cL cH ‘T
index
..I418)
cir
For bevel gears, the flash temperature index method outlined by Gleason has been widely used. Both the Gleason and AGMA methods are based on Blokand Kelley and the general form of the equation is as follows: tc = tM + tfl tc = t&f + t,
. ..(19) c 17.78
. ..(19M)
where tc is contact temperature in “F (“C); ?M is bulk temperature in OF(“C); tn
is flash temperature in “F (“C).
where %C
sac
is working contact stress number in pounds per square inch (megapascals); is allowable contact stress number in pounds per square inch (megapascals); is life factor for pitting resistance; is hardness ratio factor for pitting resistance;
cT cR
is temperature factor for pitting resistance: is reliability factor for pitting resistance.
To calculate CR , use the equations forKR shown in
8.2.7. For bevel gears, a lower coefficient of variation has been observed for contact stresses than for
8.33 Film thickness method The film thickness method is based on the work of Dowson and Higginson. The minimum oil film thickness is calculated under load and is divided by the composite surface roughness which is an average surface roughness of the pinion and gear. When the ratio of film thickness to surface roughness, h, is 1.O or greater, the surface asperities cannot touch and there is no scuffing (scoring) or wear. When the pitch line velocity of the mesh is low [PLV < 1500 ft/ min (7.6 m/s)], acceptable values of surface film thickness and h are achieved through the use of high viscosity oils. The calculations are summarized in ANSVAGMA 2001-B88 in Appendix A.
45
AGMA 91%A94
*a r= % yg cc s
6o 407
20 10 52 1 0.5
5
4
3
2
1
n
0
-1
-2
-3
Figure 16 - Reliability versus number of standard deviations
46
AGMA 911-A94
9 Gear materials and heat treatment Aerospace gears are manufactured from appropriate materials to perform under the imposed operating conditions for the life of the gear application. Gear material specifications are determined by the requirements of the application for mechanical properties, material quality, dimensional stability, hardenabilii, and manufacturing characteristics. Heat treatment is specified to achieve the required mechanical properties of the gear materials.
the material must endure. Mechanical properties of the material determine the allowable stress levels for the application.
9.2.1 Hardness The strength properties are closely related to material hardness for ferrous materials. Hardness indicates the strength of the material and resistance to tooth bending failure. Surface hardness provides resistance to gear wear, pitting and scuffing.
Ferrous and non-ferrous materials are used for aerospace gearing. Steel alloy materials are used in most primary gear applications for strength and durability. Stainless steels are used in special applications for corrosion resistance. Bronze alloys are occasionally used in worm gearing for wear resistance and reduced friction coefficient.
9.2.2 Fatigue strength
9.1 Class and grade definitions
9.2.3 Tensile strength
Aerospace gearing is divided into three classes depending on the nature of the intended application.
Tensile strength is the maximum possible tensile stress before fracture occurs. Tensile strength is considered as a mechanical property used for relative comparison of materials. Hardness is recommended in lieu of tensile strength for specifications in gear manufacture.
Gear materials are specified by grade of metallurgical control factors and allowable stress numbers according to ANSVAGMA 2006B88 and ANSVAGMA 2003-A88 The three grade numbers are: Grade lmoderate quality, Grade 2 - superior quality, and Grade 3 - Premium quality. Gear application classes are described with the material grade numbers that are used as follows: a) Typically Grade 3 material is used in main drive system components, where failure of a gear could result in the loss of the vehicle or endanger operating personnel or passengers; b) Typically Grade 2 or 3 material is used where failure of a gear may render the primary system inoperative, but where a secondary system can be engaged to perform the same function; c) Typically Grade 1 or 2 material is used where failure of a gear may affect mission capability or an auxiliary system but will not result in the loss of a vehicle nor endanger operating personnel or passengers.
Contact and bending fatigue strengths predict the number of cycles that a gear can endure at a given stress level before surface pitting or tooth fracture occurs. Contact and bending fatigue strengths are influenced by hardness, microstructure, material cleanliness, surface conditions and residual stress.
9.2.4 Yield strength Yield strength is the maximum possible stress before permanent deformation occurs.
9.2.5 Toughness Toughness is a measure of ability to absorb impact energy and is important for high impact or low temperature applications. Brittle fracture may occur in high strength materials as a sudden failure or rapid crack propagation below the tensile strength due to low fracture toughness properties. Toughness of steel gearing is adversely affected by the following factors: - Low temperature; - Improper heat treatment causing defective microstructure: - High sulphur content;
9.2 Mechanical properties
- High phosphorus and embriiling grain bound-
The particular design configuration and duty cycle
ary precipitates;
loads of the application determine stress levels that
- Non-metallic inclusions ;
47
AGMA 91%A94
- Large grain size; - Absence of nickel alloying element ; - Stress concentrating notches, fillets, tool marks or rough surface finish. 9.3 Cleanliness Cleanliness determines the extent of homogeneous material properties. Alloy steel manufactured with electric furnace practice is commonly vacuum degassed, inert atmosphere (argon) shielded and bottom poured to improve cleanliness and reduce gas content. Reduced non-metallic inclusion content improves transverse ductility and impact strength. Further refinement by vacuum arc remelting o/AR) or electroslag remelting (ESR) reduces gas and inclusion content more for improved fatigue strength. Cleanliness requirements of alloy steels are controlled according to AMS 2300 (premium aircraft quality steel) and AMS 2301 (aircraft quality steel). 9.4 Heat treatment Most wrought ferrous materials used in aerospace gearing are heat treated to meet. hardness and mechanical property requirements. Gear blanks are generally annealed and normalized to produce homogeneous microstructure for uniform machinabitii and improved response to subsequent heat treatment. Quenching and tempering increase material hardness and strength properties. Case hardening increases surface hardness and strength while maintaining a softer core for toughness.
The surface is often required to have high hardness and high strength, while the core is required to be tough for impact resistance. A correctly hardened steel gear will consist primarily of tempered martensite. Furthermore, a gas carburized gear will contain a high carbon tempered martensite case with a low carbon tempered martensite core. Excessive case carbon content can result in retained austenite if not transformed and can also lead to the formation of undesirable carbide networks at grain boundaries. Improper hardening, such as quenching, can result in undesirable bainite, free ferrite or pearlite, in the from of banding. 9.6 Hardenability Hardenability of steel is the property that determines the hardness gradient produced by quenching from the austenitizing temperature. The asquenched surface hardness is dependent primarily on the carbon content of the steel and cooling rate. The depth to which a particular hardness is achieved with a given quenching condition is a function of the hardenabilii due to carbon and alloy content of the steel. 9.7 Dimensional stability Improper heat treatment processes cause dimensional distortion and possible cracking. Distortion is caused by mechanical and thermal stresses with phase transformation during quench and tempering due to variations in: 1) section thickness, and 2) the duration at transformation temperature. 9.6 Pre-machining
9.5 Microstructure The microstructure is the material structure observed at 100X or higher magnification and reveals the constituents of the material. The constituents include, but are not limited to, martensite, ferrite, pearlite, and bainite. The microstructure also reveals grain size, carbides, carbide networks, and retained austenite. These constituents are a result of the heat treat process and can help determine if a heat treat process was done correctly or incorrectly. lt should be noted that on a cross section, the microstructure near the
surface may be dierent from the microstructureof the core. 46
stock removal
A specified minimum thickness. of surface stock should be removed from ferrous gear forgings to eliminate decarburization, seams, and other surface imperfections. 9.9 Ferrous gearing Ferrous materials for aerospace applications are primarily wrought alloy steels. Wrought steels are mechanically worked to form round stock, flat stock, or forgings. Anisotropic mechanical properties
(tensile ductilii, fatigue strength, and impact strength) vary according to the direction of hot working or inclusion and grain flow. Improved steel cleanliness improves the transverse and tangential
AGMA 911-A94
properties of forged steel that approach, but do not equal, longitudinal properties.
9.11 Material grades and heat treatment Common gear materials and heat treatment used in aerospace applications are shown in table 15.
Other ferrous materials may include H-series tool steels or austenitic, marten&c, and precipitation hardening stainless steels for special requirements. Powdered metal materials may be used in certain non-critical applications.
9.12 Gear surface hardening
9.10 Non-ferrous gearing Copper base (bronze) gears are used in worm gear applications with steel worms to improve wear resistance and reduce the coefficient of sliding friction. Manganese bronze and aluminum bronze have higher strength than phosphor or tin bronze materials.
Most aerospace power gears are produced with heat treated alloy steels and surface hardened to provide tooth bending strength and resistance to pitting and wear. Carburizing is the primary method of surface hardening gear teeth. Nitriding is an alternate surface hardening process specified where carburizing and quenching would cause excessive distortion.
Table 15 - Typical aerospace gear materials
Material
AMS SPec
Heat treatment*
Typical hardness Case, HRC**
Core, HRC
Typical applications
Alloy St-L . AISI 9310
626516260 C-H
58-62
3442
58-62 58-62
4248 3844
58-60 59-64
48-52 3642
58-62 58-62 60-64 60-64 48-53
35-45 3442 3442 38-44 27-35
Main drive, Accessory, actuators Actuators Main drive, High temperature Actuators Main driie, High temperature High temperature High temperature Accessory drive Accessory drive Accessory drive Accessory drive Actuators
4330M VASCO X2M”’
6427 N/A
HP 9-4-30 PYROWEAR 53***
6526 6308
C-H
M50NiL CBS600 Nitralloy 135M Niiralloy N ‘AISI 4340 AISI 4340 300M Stainless steel: PH13-8 MO Custom 455
6278 6255 6471 6475 6414 6414 6419
C-H GH TH-N TH-N TH-N TH TH
5629 5617
PH ST
Accessory drive Actuators
ST
Worm gear Actuators
Bronze: C63000 NOTES-
4640
C-H = Carburize and harden TH-N = Through harden and nitride TH 5: Through harden = Precipitation harden E = Solutionheat treat
52-55
(1OOHRB)
*
Rockwell hardness scales (HRC and HRB) are shown for direct comparison only. In general, those scales are not specifically recommended for measurement where other, more accurate hardness scales are commonly used.
-*
Proprietary material designation.
49
AGMA 911-A94
However, nitriding does not increase tooth bending strength as much as carburizing. Other methods of selective direct hardening may be used for less critical gear applications with medium carbon alloy steels. The minimum depth of surface hardened case in gear teeth is required to resist: (1) sub-surface shear stress caused by tooth contact pressure and (2) root fillet tensile stress caused by tooth bending. Minimum required case depth is typically specified as the greater value of: (1) two times (2x) the depth of maximum sub-surface shear stress at the pitch line, or (2) 10 percent of the tooth thickness to accommodate sub-surface fillet tensile stress. Refer to ANSVAGMA 2001-B88 for carburiied or nitrided case depth recommendations. The case must not be so great as to result in brittle teeth tips and edges, or high residual tensile stresses in the core. Maximum case depth at the tooth tip should be limited to 56 percent of the tooth top land thickness when possible. The effective case depth for carburized and hardened gears is defined as the depth below the surface at which the hardness is 50 HRC equivalent by conversion from microhardness. The effective case depth for induction hardened gears is defined as the depth below the surface at which the hardness is 10 HRC points below the specified minimum surface hardness. For nitrided gears, case hardness is specified as total case depth, and is defined as the depth below the surface at which the hardness has dropped to 110 percent of the core hardness. The practical limit for maximum nitrided case depth is 0.040 inch (1 mm).
After carburizing the gear is cooled slowly in the carburizing medium or protective atmosphere to control decarburiiation. Next the gear may be reheated to 1200°F (649OC)for subcriiical anneal to reduce sudden changes in carbon content across grain boundaries in the microstructure. After cooling again, the gear is reheated to 1500°F (816°C) to austenitize the carburized steel. When the gear is quenched in oil, much of the austenite transforms into hard and brittle martensite. A deep freeze cycle may be used to complete the transformation and minimize retained austenite. Finally the gear is tempered at 300-360°F (149-182OC) to reduce brittleness and microcracking. The carburizing and hardening cycles produce a very hard, mattensitic layer on the surface with a less hard, tough core. In addition, the carburiiing and quenching processes cause high residual compressive stresses in the surface for increased material strength and resistance to fatigue failure.
9.12.1.1 Carburizing process control Gas carburizing atmospheres can be in equilibrium with a wide variety of carbon content. Proper surface carbon content must be maintained to achieve the required properties after heat treatment. Surface carbon content should be maintained between 0.6% carbon and 1.0% carbon. Less than 0.6% surface carbon will lead to low surface hardness which may not meet final requirements. Surface carbon contents over 1 .O% may lead to excessive retained austenite and/or grain boundary carbide networks. With higher surface carbon contents it may be necessary to sub-zero cool to transform retained austenite to martensite.
In a typical process for alloy steel, gas carburiiing diffuses carbon into steel from a hydrocarbon gas atmosphere while the gear is heated to 1660°F (899OC)or higher in a furnace. Carbon diision into the steel is affected by (1) timeof exposure, (2) temperature in the carburizing atmosphere, and (3) percentage of carbon in the atmosphere. The higher
The gas carburizing atmosphere is controlled to obtain the desired carbon content in the surface of the gear. It is diicult to measure carbon potential directly. Cften one or more components of the gas mixture, such as CO,, CH4, H20, or 02 are measured and the other component (usually CO) is assumed to be in proportion. However, the unmeasured component may vary due to variable gas source, air leak or carbon build up. If CO, hydrogen
concentration of carbon near the surface provides
or other componentsvary,the reactionswill shii and
maximum local hardness.
make the carbon potential relationship inaccurate.
9.12.1 Carburizing
50
AGMA Qll-A94
Shim stock (a thin, flat material sample) can be used to accurately assess the initial carbon potential of the carbutizing gas. However, conditions may be different later when the gears are exposed in the furnace. 9.12.1.2 Other carburizing methods Vacuum carburizing and plasma carburizing produce results similar to gas carburiiing. Vacuum carburizing is often faster because vacuum carbuiiing is carried out at higher temperatures [I 80& 1950°F (982-l 066 “C)], rather than the 1650-l 800°F (899-982 “C) typically used ip gas carburizing. Plasma carburiiing occurs in a DC electrically charged furnace with the gears acting as the cathode. A plasma is formed around the gears which enhances the absorption of carbon at the carburizing surface of the gear. 9.12.2 Nitriding Nitriding is an alternate case hardening process often specified for gears when distortion would be difficult to control if the gears were case carburized and quenched. In nitriding, nitrogen is introduced into the surface of the steel at relatively low temperature [925-1050°F (496-566 “C)] from a nitrogen containing atmosphere such as ammonia. A hard case is produced by the formation of hard nitride compounds in the surface making quenching unnecessary. Special steels are needed that contain elements such as aluminum or chromium to form hard nitrides during treatment. The steel is nitrided in the hardened and tempered condition. The process is controlled by adjusting the dissociation of the ammonia. More often the Floe process is used which is a double-stage process analogous to the boostdiffuse cycle in carburizing. In the first stage, dissociation level is controlled at 1530% by using a temperature range of 925- 975”F( 496-524 “C) producing a white nitride layer which is diised in the secpnd stage by increasing the dissociation to 80-85%. The high dissociation can be achieved by increasing the temperature to 1025-l 050°F (552-566 “C) and using an external dissociation. The remaining white layer should be limited to a maximum allowable depth of 0.0005 inch (0.013
mm). Some critical applications require removal of the white layer. Even with the two-stage process, nitriding is slow, taking about ten times as long as carburizing to produce a specified case. The nitride process produces a very hard case with minimum distortion. Volume increases during nitriding and causes favorable compressive stress to build up in the case. However, nitrides tend to accumulate more at tooth edges. 9.12.2.1 ion nitriding Ion nitriding or plasma nitriding is similar to plasma carbutizing in that a plasma is formed around the work during treatment. 9.12.3 Selective direct hardening Selective direct hardening produces a hard case by heating the surface layer above the austenitiiing temperature and rapidly quenching, while leaving the core in the original condition. A medium carbon steel is used with the required carbon already in the steel and heat treated to proper core hardness. Since a large proportion of the part remains cool, thus stabilizing the material, distortion is much less than it would be if the entire part were heated. The higher carbon containing material may make machining more diicult than a carbuiiing grade of steel. Induction hardening is the one method of selective directive hardening that may be suitable for some production aerospace gears. However, the induction hardening process should be developed to control residual stresses and annealing between the case and core. Two other surface hardening processes, laser heat treatment and electron beam heat treatment, are also being developed. 9.12.3.1 Induction hardening Induction hardening is achieved by using an alternating current in a work coil that surrounds the part to be heated. An alternating magnetic field is established that induces a potential in the part causing a current to flow in the closed circuit. Heating is produced by the resistance to the induced current. The rate of heating depends on the strength of the magnetic field. The depth of the field varies inversely with the frequency of alternation. The 51
AGMA 91%A94
higher the frequency, the more shallow the heating effect. If a circular coil is used to heat a gear then the tips of the gear are coupled closer to the coil and thus they heat more, resulting in a deeper case depth at the tooth tips. After the heating is complete, the current is turned off and the part is quenched by synchronized jets of a quenching fluid, usually water-based. 9.12.3.2 Laser heat treatment Laser heat treatment is a surface-hardening process in which laser energy heats the surface above the austenitiiing temperature. When the source of energy is removed, the part self-quenches by diffusion of the heat into the mass of the part. The laser causes a steep temperature gradient due to the extremely rapid heating rate. As the rate of heat input increases, the depth of hardening is reduced, since the temperature gradient becomes steeper and the surface temperature must be limited to avoid melting. To spread the laser energy over required coverage area, the beam is usually defocused. Alternatively oscillating optics or integration optics with a faceted mirror may be used. Laser beam penetration is controlled by power level and rate of beam traverse. Penetration increases with increased power and decreases with increased beam traverse rate. 9.12.3.3 Electron beam heat treating This method is similar in principle to laser heat treating, except that heating is achieved by an accelerated stream of electrons instead of a light or infrared beam. When the electron beam is turned off, the part self-quenches. The electron beam heat treat process occurs in a vacuum environment. This requirement introduces some complications into the f&king. The electron beam is manipulated by magnetic coils. 9.12.4 Residual stress patterns One advantage of case carburized parts is that when the treatment is properly carried out it produces compressive stress at the surface. Compressive
52
stresses help counteract tensile stresses produced during tooth loading and thus increase the expected life. A part that has been carburized is heated above the austenitizing temperature and then quenched. The surface cools faster than the center of the section because heat is abstracted from the surface by the quenching media. The net result is that transformation of austenite to martensite starts at the case/core interface with an expansion as martensite is formed. The case is the last material to transform and expand to martensite, causing compressive stresses because the core has already transformed and restrains the case. The conditions are different in selective hardening, but the results are similar. Energy is transmitted quickly into the surface resulting in a surface layer heated above the austenitizing temperature. This layer will later become the hardened case. When the energy is turned off, rapid cooling progresses and again the case is the last to transform and the restraint induces residual compressive stresses as the surface expands during transformation from austenite to martensite. 9.12.5 Dimensional problems caused by heat treatment Heat treatment tends to cause more quality problems than any other manufacturing step. Heat treatment causes dimensional changes due to volume change resulting from phase transformation. Distortion occurs from a combination of aeometric factors and uncontrolled stress relief. These two factors acting together often cause unpredictable results. Variables that contribute towards the dimensional changes include: - Variations in material composition; - Residual stress differences; - Size of part (within tolerance range) before heat treatment; - Surface condition; - Cart&zing heating cycle; - Carburizing atmosphere control; - Depth of case; - Quenching parameters; - Quenching die dimensions; - Post heat treatment.
AGMA 9ll-A94
Gear manufacturers try to bring the component size under control in the finish grinding stage. If excess material is left on the part prior to heat treatment, there will be enough stock to enable the size to be brought under control. However, if too much is taken off, the most effective portions of the carburiied (or nitrided) case are removed. Figure 17 shows uniform material being removed from a tooth after heat treatment. In the example shown in figure 18, the tooth has distorted to the right. To correct the profile, excess stock has to be ground from the right side of the tooth. This has several serious consequences. Size before Heat Treatment
’ Case Figure 17 - Schematic of material ground from a gear tooth m-h
Tooth Distortion Heat Treatment
First, there is lackof uniformity in case depth leading to uneven residual stress distribution. Second (and worse) is that the gear appears satisfactory in a nondestructive inspection, even though the performance of the gear will be less than optimum. Third, a considerable thickness of material has to be removed during grinding, increasing the probability of grinding bums. Some problems that are blamed on grinding can in reality be traced back to heat treatment. Thus the effects of heat treatment have to be considered before and after the process in both the soft machining and hard finishing stages. Selective direct hardening processes minimize distortion and associated rework problems. 9.13 Gear through hardening Some aerospace gears that are not critically loaded may be heat treated by through hardening processes to obtain the required mechanical properties for the gear application. Typical material hardnesses range from 32 HRC to 54 HRC with allowable gear tooth bending and contact stresses that are lower than those for surface hardened gears. The process includes annealing, normalizing, quenching and tempering. For annealing, ferrous alloys are heated to 1475-l 650°F (802-899%) and furnace cooled below 600°F (316%). In normalizing, ferrous alloys are heated to 1600-l 800°F (871-982°C) and cooled in air. Annealing and normalizing are used as homogenizing pm-heat treatment processes to reduce metallurgical nonuniformities. For the quench and temper process, ferrous alloys are heated to 1475-16OOOF(802-871 “C) to form austenite, followed by rapid quenching which causes the gear to become harder and stronger by formation of martensite. The gear is then tempered to a specific temperature below 1275OF (691 “C) to achieve the desired hardness and strength with improved ductility and toughness.
Figure 18 - Schematic of material ground from a distorted gear tooth
53
AGMA 91%A94
IO Surface treatment This information covers a variety of currently utilized surface treatments, generally applied after hardening, that are used to enhance the durability of aerospace gears with respect to their resistance to metal fatigue, wear, and environmental corrosion.
10.1 Introduction Post hardening surface treatments are usually employed on gearing to accomplish one or more of the following: - Raise the bending fatigue strength at the tooth root fillet radii, as well as in the rim, web, hub and integral shafts;
‘B Y
L1oo 53 ts : .g 80 8
- Increase the contact fatigue properties along the tooth flanks near the pitch line; - Improve the resistance of gears, bearings, and other mating surfaces to adhesive wear phenomenasuch asfretting, scuffing and galling; - Improve the resistance to abrasive wear, such as scuffing (scoring) from ingested abrasive and wear debris; - Improve lubricity, both for break-in and continuous operation; - Increase resistance to general and to localized corrosion (pitting, stress corrosion, etc.) during storage, from aggressive service environments, or due to contaminated lubricants.
40 -80
0
80
160
tension compression Peak residual stress, ksi
Figure 19 - Fatigue strength in ground AISI 4340 (50 HRC) [ll] Tensile residual stresses may be present at the surface of gears after heat treatment, particularly in high strength, through hardened steels tempered at temperatures which provide only moderate relief of quenching stresses. Other processes which may also introduce residual tensile stresses in gears include grinding, abusive machining, hard plating (such as nickel and chromium) and hard thermal Some of the non-traditional spray coatings. machining processes such as electro-discharge machining (EDM) and electrochemical machining (ECM) may also damage surfaces through the formulation of a brittle recast layer by EDM or the introduction of pitting and/or shallow intergranular attack by ECM.
Tooth bending fatigue is of primary importance since fatigue fractures of gear teeth can result in catastrophic destruction of a gear train. Contact fatigue pitting, wear, and corrosion, however, can also lead to tooth fatigue fractures from a reduction in load carrying area and/or through the formulation of local stress raisers. Since most geartooth fatigue cracking initiates at or very near the surface, fatigue properties are highly sensitive to surface characteristics such as roughness, and to residual surface stresses introduced during the manufacturing process. As an example shown in figure 19, surface residual tensile stresses can lower fatigue strength,
On the other hand, surface hardening heat treatments such as carburizing and nitriding introduce beneficial residual compressive stresses in gears. Some of their benefits may be offset, however, by the presence of shallow, partial decarburization and intergranular oxidation in unground roots of some
while fatigue strength is raised by the presence of
carburized gears, and the presence of a brittle white
surface residual compressive stresses.
nitride layer in some nitrided gears. Honing, polish-
54
AGMA 911-A94
ing, burnishing and other “superfinishing” techniques are often eff e&e in removing fatigue lowering surface layers and defects, as well as in reducing surface roughness. Another effective process for improving fatigue properties by inducing high surface residual compressive stresses in gears is “controlled shot The process is utilized for both peening.” carburized case hardened and through hardened aerospace gears. Controlled shot peening should not be confused with “shot blasting,” “grit blasting,” or “abras.ive blasting.” The latter processes are employed for cleaning or abrading surfaces and do not produce consistent residual surface compressave stress profiles or predictable increases in gear fatigue properties. 10.2 Shot peening Controlled shot peening is a surface cold working process in which hard, spherical shaped media (steel, ceramic or glass) are propelled at relatively high velocity and at a nearly normal incidence angle against a workpiece. Its purpose is to promote surface strain hardening and to induce predictable surface and near surface residual compressive stresses. Each particle of round shot striking the surface acts like a tiny peening hammer, producing a small indentation or dimple. The surface fibers are stretched (yielded in tension) by the dimple formation, which also forms a sub-surface hemisphere of strained metal below the dimple. Overlapping dimples develop an even layer of plastically deformed surface fibers and a sut+sutface zone of strained material. If unrestrained by the core, the surface of the workpiece would elongate under these deformation induced strains. The greater mass of unaffected metal in the core, however, restricts this expansion, producing high magnitude residual compressive stresses in the surface and near surface layers.
forces between surface and core are balanced. Since the offsetting tensile stresses act over a significantly greater cross-section than the compressive stresses, the tensile stresses are generally of low magnitude. See figure 20. In shot ieening thin sections, the depth of peening is controlled so as to keep the core tensile stresses at moderate values and also to prevent distortion of the workpiece. Compression Tension % Ultimate tensile strength (-) (+) -100 -50 40 0 +1 00 I I 0 t ss-l -28 IE -4: 4 wnax 6 - 6$ L -6nax -8% 25 28 - 12 SS is surface stress CS is compressive stress TS is tensile stress Figure 20 - Example of residual stress profile created by shot peening [12] The maximum residual compressive stresses from shot peening are at least as great as half the yield strength of the workpiece material, providing that the media used for peening has a minimum hardness at least as hard as the workpiece. For example, the use of regular hardness steel shot (HRC 45-52) will not induce the same magnitude of compressive stress in a carburized gear as special hardness shot (HRC 55-62). This is illustrated in figures 21 and 22.
The maximum compressive stress is generally located just below the surface and decreases with increasing depth. The depth at which the residual compressive stress becomes zero is usually referred to as the effective depth of peening. Beyond
The stress profiles in these and subsequent figures were determined by standard X-ray diffraction techniques for measuring residual stress (SAE J784a). Subsurface measurements are made after elecropolishing away surface layers and are corrected for subsurface stress gradients [20]. Since the use of the harder shot increases surface roughness, it may be necessary to final finish after
this point, the compressive stresses are offset by
peening when surface finish is critical. Lapping and
sub-surface
honing
residual tensile stresses so that the
may be used if the operation
does not
55 55
AGMA 911-A94
remove more than 10% of the compressive stress depth. Regular hardness shot without post peen finishing may also be used; however, the maximum compressive stresses will be lower than that obtainable with the harder shot. .
0
‘Y5
/-
G 5 - -500 - -1000
I 0
I
0.004 0.008 Depth in inches
0.012
Figure 21 - Peening 1045 steel at 46 HRC with 330 shot [13]
0.004 0.008 Depth in inches
Figure 22 - Peening 1046 steel at 62 HRC with 330 shot [13] 10.2.1 Crack arrest due to shot peening Metal fatigue cracking takes place in two distinct phases. The initiation phase encompasses the development and early growth of a small crack, almost always at a “free surface.” The propagation phase is that portion of the total life during which the crack grows to the point of failure. It is often difficult to define the transition from initiation to propagation. In assessing the effect of residual compressive stresses in increasing fatigue strength and life, it is usually simpler to consider fatigue from the standpoint of fracture mechanics theory. This assumes
that minute cracks or crack-like flaws are always present in engineering structures and that fatigue
56
life is primarily dependent on the propagation or non-propagation of these cracks, which in turn depends on the stress conditions at the crack tip. The theory of crack arrest due to the residual compressive stresses induced by shot peening is based on the following: - A crack will not propagate unless a tensile stress forces it open near the crack tip; - The crack tip will not open as long as a compressive force acts upon it. Some important points to be considered with respect to shot peening are: -Gears should always be cleaned and inspected thoroughly for cracks before shot peening. If cracksarefound, the cause should be thoroughly investigated and corrected. If allowed, the parts should then be repaired by blending or re-machining. Shot peening tends to obscure cracks and should not be used as a method for repairing -Shot peening is effective as long as the residual compressive stresses do not fade out or relax due to exposure to high temperatures (generally 500°F max. for most steels) or to over-stressing (applied stresses in excess of the yield strength); - Shot peening effectiveness is greatest when to an area of a part where there is an applied stress gradient, such as in bending or torsion, or in the presence of geometric notches, such as fillet radii; -Shot peening is more effective in the high cycle fatigue (HCF) domain; however, it is also beneficial in the low cycle fatigue (LCF) region up to the point where the applied stresses correspond to the net section yield strength. 10.22 Effect of shot peening on tooth bending fatigue The tooth root fillet radii in aerospace gears are generally subject to relatively high cyclic bending stresses. Carburized case hardened gears, as well as through hardened gears, are frequently shot peened in the root and root fillet area to introduce a high magnitude compressive stress, as illustrated in figure 23. Significant increases in bending fatigue strength can be realized when shot peening
carburiied and hardened gears with unground roots, as well as in gears with ground rcotswhen the peening is performedafter grinding. See figure 24.
AGMA 911-A94
Non-cased areas in carburized gears, such as the fillet radii at rims, webs, hubsand shafts, also exhibit improved fatigue properties after shot peening. See figures 25 and 26. NOTE - As bending stresses approach the yield strength of the material, the effectiveness of shot peening is reduced due to plastic deformation which causes fading of the residual stresses.
6Ir” ‘T 5 t 5 t E4 I 8 P a3 d 0 z2 0 Y z! n1 0
I
I
Spiral bevel gear, 3.625 DP Material, 8620H bar stock, heat treated: Carburizsd & hardened to 61 HRC Load bansmitt& 135 HP @ 5000 RPM Shot peeningspecifications: Shot size 170 H 010 -014A Intensity Coverage 200%
.I
104
165
lb6
107
Cycles to failure Figure 24 - Increase in fatigue resistance of spiral bevel gear 1151 Figure 23 - Stress profile of carburized gear tooth root, ground and then shot peened with special hardness shot [14] 60
420
l
350 ‘G y-40, i!
4 I
zz 280 nl
Shot peened - 43 ksi (300 N/mm2)
I
I
I
“::;f:; 1%
5x106 107 5x105 106 Revolutions
I
\ .
II
I
YFymnl(o(0124”)
I
I
1 1
5x107 Cycles tti failure
Figure 25 - Fatigue tests on rear axle shafts [16]
Figure 26 - Fatigue tests on notched shafts [17]
57
AGMA 911-A94
109.3 Effect of shot peening on tooth contact fatigue
investigator has also indicated that this lower friction can also reduce gear noise.
Contact fatigue pitting and spalling in gear teeth generally initiates at the dedendum surface of the tooth flank, just below the pitch line. Surface initiation results from the combined effect of subsurface shear stresses from rolling and tensile frictional traction stresses from negative sliding (rolling and sliding in opposite directions). Testing of peened and unpeened carburized and hardened gears has indicated that shot peening of the tooth flanks increases contact fatigue life by a factor of about 1.6. See figure 27.
The modem “delamination theory” of wear suggests that all adhesive-wear processes are related to fatigue and that the residual compressive surface stresses from shot peening, as a result, may also increase resistance to wear. This has not as yet been confirmed in any rigorous test program. 10.2.5 Effect of shot peening on corrosion In gear steels, shot peening has no effect on either Shot general corrosion or pitting corrosion. peening, however, is effective in retarding the initiation of fatigue cracking from corrosion pits until the depth of the pits exceeds the effective compressive depth of the peening.
10.2.4 Effect of shot peening on wear High hardness gear steels exhibit only low to moderate strain hardening when shot peened. The increase in surface hardness due to shot peening, therefore, has little effect on increasing resistance to adhesive or abrasive wear. The tiny indentations on the gear flanks produced by peening, however, act as very small oil reservoirs which can help to promote better lubrication, thereby reducing fretting and scuffing (scoring), as well as lowering operating temperatures by reducing friction. At least one 99
1
I
I
I
I
I
I
Several of the ultra-high strength, through hardened gear steels are susceptible to stress corrosion cracking (SCC) when exposed to a moist chloride atmosphere. The residual surface compressive stresses from shot peening, often in combination with a sacrificial coating (such as cadmium, aluminum and zinc) may prevent or retard SCC in these steels as well as fatigue cracks initiated by SCC.
1 -8 z
95
s
a0
E z $ t & 3
60
ii E 10
E 0.
6 4
20
100
40
Number of cycles xl O6 Standard gears Comparison of surface (pitting) fatigue lives of standard ground and shot-peened carburized and hardened CVM AISI 9310 steel spur gears:
20
I
I
I 100
Number of cycles xl O6 Shot-peened
gears
Speed - 10 000 RPM Lubricant - synthetic paraffinic oil Gear temperature - 77°C (170°F)
Figure 27 - Fatigue life comparison 1181
58
I 40
I 200
AGMA 91%A94
10.2.6 Shot peening process controls There is presently no non-destructive production method to determine the proper shot peening of a gear. Strict control of the shot peening process is, therefore, essential to ensure repeatability and uniformity on a part-to-part and lot-to-lot basic, while conforming to applicable specifications. The shot material, size, shape and hardness, as well as velocity and impact angle, must be rigidly controlled to provide consistency in peening results. In order to effectively control the shot peening process, the following parameters must be addressed: - Media control - Intensity control; - Coverage control; - Equipment control. Details can be found in shot peening specifications such as MIL-S-1316X, AMS2430, AMS2431 and AMS2432, as well as the SAE Manual of Shot Peening, J808a. The following, therefore, is meant as only a brief summary. 10.2.6.1 Media control Media typically used for shot peening are small spheres of cast steel, conditioned cut wire (both carbon and stainless steel), ceramic and glass materials. Peening media must be uniform in size and essentially spherical in shape with no sharp edges or broken particles. Broken or sharp edge particles can be potentially damaging to the part surface. 10.2.6.2 intensity control Calibration of the impact energy orpeening intensity of the shot stream is essential to controlled shot peening. The energy of the shot stream is a function
The height of this curvature when measured in a standard Almen gauge is called arc height. There are three standard Almen strips currently in use: the “A” strip, 0.051 in (1.3 mm) thick for intermediate intensities; the “C” strip, 0.094 in (2.4 mm) thick for high intensities; and the “N”strip, 0.031 in (0.79 mm) thick for low intensities. Three strip thicknesses are required since the useable range of curvature on the Almen strips is 0.004 to 0.024 inches (0.1-0.6 mm). A comparison of the intensities of the A strip for the C strip and the N strip, as indicated by arc heights, is also shown. See figure 28. Intensity designations must include both the arc height and type of strip used. Substitution between strips is not permitted. An Almen arc height is not properly termed intensity unless “saturation” is achieved. This is done by developing an intensity saturation curve. Saturation is defined as the earliest point on the curve where doubling the exposure time produces no more than a 10% increase in arc height. Most important to the user is the fact that the depth of the compressive layer is proportional to the Almen intensity.
inch
0.001 2
of the media size, material, hardness, velocity and impingement angle. In order to specify, measure and calibrate intensity, a method developed by J. 0. Almen utilizing SAE 1070 spring steel specimens called “Almen strips” is still in use. The unpeened strip is fastened to a steel block and exposed to a steam of peening shot for a given period of time. Upon removal from the block, the residual compressive stress and surface plastic deformation produced by the peening impacts cause the Almen strip to curve, convex on the peened surface.
4
6
8 130
4
8
12
0.001 inch 16
20
0.30
0.20
- 10 r .-E 0
0.05
0.10
0.15
0.20
0.10 E
B 6
Almen intensity of C strip, mm Figure 28 - Correlation of Aimen intensities as indicated by arc heights of A, C, and N strips peened under identical blast and exposure conditions.
59
AGMA 91%A94
10.2.6.3 Coverage control Coverage is defined as the extent, in percent, of uniform and complete dimpling or obliteration of the original workpiece surface. Inspection of percent coverage can be accomplished by using a ten power magnifying lens or through the use of an approved fluorescent liquid tracer system. It is extremely difficult to visually determine coverage in hardened gears with a ten power visual examination. For determining coverage on gears, the liquid fluorescent tracer system is widely used. The coating, which is applied prior to peening, iS removed during peening at a rate proportional to the percent of shot peening coverage. Examination under black light (UV) provides a practical and superior method of verifying the 100% minimum coverage required when shot peening gears for fatigue resistance. 102.6.4 Equipment control
When shot peening is selected for man-flight vehicle geared components, for geared components where it is used as part of the design strength, or for geared components which are considered critical to system success, the following are suggested: - Shot peening by computer monitored/controlled equipment per an applicable, approved specification such as AMS2432, including the use of liquid tracer systems to verify coverage; - Prior approval of a strength and life analysis which justifies the approach. This includes life certification/qualification testing on hardware conforming to the production configuration; -The establishment and validation of an inspection interval that takes into account the potential for in service degradation of peening benefits; - Any repair/rework peening required at prescribed inspection intervals to attain the required fatigue life.
The machines used for shot peening provide means for propelling shot by air pressure or centrifugal force against the workpiece, as well as mechanical systems for moving the work through the shot stream or moving the shot stream through the work by translation and/or rotation. The equipment also continuously removes broken or defective shot so that it is not used for peening. Many modem shot peening machines are computer controlled. Typical peening parameters which are monitored, controlled and documented are:
10.2.8 Guidelines for media and intensity selection
- Air pressure of each nozzle or wheel speed of each wheel;
- Requirements for peening internal surfaces and/or intersecting holes.
- Shot flow of each nozzle or wheel; - Part rotation or translation rates; - Nozzle reciprocation rates, distances and run times; - Total cycle time.
The proper selection of media type, size, hardness and intensities for peening of aerospace gearing is dependent on numerous variable including: - Minimum fillet radii in gear; - Thickness of peened section of gear and/or effective depth of carburized case, if any; - Hardness of gear in areas to be peened; - Surface finish requirements;
The tables of guidelines found in numerous specifications, such as MIL-S-13165C, are valuable statting points for selection. NOTE - Magnetic particle, penetrant, ultrasonic or otherflaw andcrack detectionmethodsshouldbe performedprior to shot peening.
102.7 Shot peening design considerations
10.3 Surface coatings
When the gear user does not wish to take advantage of the fatigue strength increases in the design calculations, but wishes to use the process to overcome many of the residual surface tensile stresses or similar problems from previous manufacturing
Table 16 lists of some of the most widely used coatings for aerospace gears. The table is divided into three broad categories of coating application: antifretting barriers, corrosion resistance, and build-up repair.
processes, shot peening controlled per AMS2430
Under the anti-fretting barrier list are coatings used
or the equivalent is generally suitable.
to reduce the effects of fretting at joints. Aluminum
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AGMA Qll-A94
bronze econal has been successfully used between mounting surfaces of gears which are bolted to flanges. Silver and copper plate have also been used on gear bolted connections, but are more normally used as coatings for splines. Tungsten carbide has had application to cone seat connections between helicopter rotor heads and main rotor shafts under high loads and stresses. Steel parts exposed to the atmosphere have used baked resin (or other forms of paint) in non-working areas for corrosion resistance. Cadmium plate has been used on steel parts exposed to the atmosphere with brush cadmium being used on high strength steel (over 150 ksi) and conventional cadmium on lower strength components. Thin dense chrome has been used on grease lubricated bearings exposed to the atmosphere. There is disagreement over the benefits vs. cost of black oxide or phosphate coatings on steel parts. The third broad category of coatings used in aerospace applications is in the area of repair of surfaces that are under the minimum material condition. Sulfamate nickel plate has been used as a repair for bearing journals that are up to 0.010 inch undersized, in carburized and uncarburized conditions. Since there may be a reduction in fatigue strength from sulfamate nickel plating, repairs should be made in areas of low stress only. Electroless nickel and brush nickel plate have been used for very thin coatings where precision is required. The parts need not be ground after plating as the thickness is uniform and the plating takes the form and finish of the base material.
fields. The ion implantation process is primarily used for two reasons: - Improving surface hardness, wear resistance against adhesive, abrasive and scuffing conditions; - For improving corrosion materials.
resistance
of
In the ion implantation process, energetic ions impinge on the surface of interest. The ions penetrate into a substrate material, modifying the surface by changing crystallinity and chemical structure of the material. Ion implantation has proven to be an effective and technically attractive approach for changing the surface properties of high value added and high precision components of various criiical systems. Ion implantation of gears was introduced and has been used in the last ten years for a variety of reasons, primarily for improving the mechanical performance in a wide range of applications and improving corrosion resistance in helicopter transmissions and aircraft engines. 10.42 Improved mechanical performance of the mr Ion implantation has been used on a variety of gears for improving surface hardness, providing lubrication and minimizing scuffing wear of gears. The most common material studied in this field is AISI 9310 steel, a commonly used material for gear fabrication.
10.4.1 Introduction
Ion implantation is also used extensively for treatment of titanium alloy gears. In aerospace applications, for weight saving purposes, titanium gears are occasionally used. Ion implantation of ion species, such as nitrogen and/or carbon, into titanium alloys will induce hard phase precipitates of titanium nitride and/or titanium carbide. The increased surface hardness of the gears and lower coefficient of friction induced by ion implantation has proven very effective in reducing scuffing wear and lowering wear of titanium components.
Ion implantation isan effective processforchanging the surface properties of materials without adversely affecting the bulk properties. Ion implantation technology is used in a variety of applications in aerospace, automotive, cutting tool, biomaterials, metal stamping and metal piercing
In hydraulic fuel applications, ion implantation has been used to modify the surface properties of the face of the gear. In this case, the main concern is excessively high friction, adhesive wear and particle debris generation. ion implantation of titanium and carbon into commonly used steels used in fuel
For any coating that is applied in areas that are clamped or otherwise in contact, consideration must be given to the consequence of loss of coating thickness, loss of preload on bolted joints, loss of fits for press fitted parts, etc. 10.4 Ion implantation of gears
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pumps has proven very effective in minimizing friction and improving resistance of the surface against wear. 10.4.3 Ion implantation for corrosion resistance ion implantation has been used extensively to modify corrosion resistance of steel gears. The most common recipes include ion species such as chromium, molybdenum, tantalum and titanium followed by carbon ion implantation. Depending on the nature of corrosive attack, one or the other of these ion species have been studied in laboratory
testing, as well as field appiications. Implantation of ion species such as chromium and molybdenum have significantly improved pitting corrosion and aqueous corrosion in salt bath applications. 10.4.4 Conclusion High precision gears appear to be ideally suited for the ion implantation process. The available data supports the use of the technology on a broader range of applications requiring superior corrosion and wear resistance.
Table 16 - Surface coatings used in aerospace gear units Coating
Application
Specification
Ant--fretting Aluminum bronze econal Copper Silver plate Sulfamate nickel plate Tungsten carbide Detonation thermal spray Tungsten carbide plasma spray Corrosion resistance
Fretting barrier (plasma spray) Fretting barrier Fretting barrier (SplinesIJoints) Limited fretting barrier Hard, wear resistant, fretting barrier
METCO 604NS MIL-C-14550 QQ-S-365 QQ-N-290 None
Hard, wear resistant, fretting barrier
METCO 072VF
Baked resin Black oxide Brush cadmium plate
Corrosion resistance Corrosion resistance (on shelf) Corrosion resistance. Used on high strength parts Corrosion resistance for parts outside gearbox Touch up for Black Oxide Corrosion resistance Corrosion resistance & break in surface for gears Corrosion resistance, hard wear surface
MIL-R-3043 MIL-C-13924 MIL-STD-665
Thin build up repair (0.0002 inches thick). Precision plate with no grinding. Thin build up repair (0.0002 inches thick) No machining after plating. Uniform plating thickness even in grooves. Loss of fatigue strength. Build up repair (0.005 inches thick)
MlL-STD-665
Cadmium plate Gun blue Ion vapor deposition Phosphate coat Thin dense chrome
QQ-P-416 MIL-C-13924 MlL-C-6637 MIL-P-16232 QQ-C-320
Build-uo reoair Brush nickel plate Electroless nickel plate
Sulfamate nickel plate
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QQ-N-290
AGMA 911-A94
11 Manufacturing
considerations
11.I Introduction The performance of an aerospace transmission is directly related to the manufacturing process used to fabricate the gears. Due to the unique characteristics of the sliding mesh, close control of the gear tooth geometry is paramount to yield a successful application. The metallurgical and geometric quality must be controlled to meet the drawing requirement. In addition, surface finish, including surface roughness and waviness, is important to reduce surface distress. The Engineering drawing requirements should dictate the manufacturing source and process used to fabricate the gears. The selection process for the manufacturer should include the approval of Engineering, Manufacturing Engineering, and Quality Assurance. The manufacturer should provide tangible evidence of their capability to meet the drawing requirements.
11.2 Spur and helical gears Manufacturing spur and helical gears require unique machines to cut (generate) the gear teeth prior to the carburizing and hardening process. Generally, the parallel axis gear teeth are semi-finished, using a shaping or hobbing machine. The shaping or hobbing cutters should be designed to accommodate the appropriate stock on the gear teeth to meet the needs to clean up the gear teeth after hardening. Care should be taken not to leave too much stockon the teeth, which would reduce the effective case depth and hardness. The tooth form (involute and lead) and tooth accuracy (pitch and index variation) should be controlled within reasonable tolerances to assure proper stock allowance for final grinding of the teeth. Sometimes, 200 percent of the drawing tolerance for the above requirements can be used to control the as-cut dimensions. The actual amount needs to be carefully chosen based on the material and configuration.
11.2.1 Shaping gears The shaping machine process can be used to generate internal and external gears, especially gears which have a shoulder or other physical obstruction limiting the path of thecutter. Shaping machinesare unique in that they can generate most or all parallel axis gearswhen properly tooled and fuctured. Shap ing helical gears will require a special helical guide.
Special shaping machines are available to shape, crown, and/or taper the spur or helical tooth. The involute profjle is a function of the involute shaping cutter and the generating machine motion.
11.2.2 Hobbing gears Hobbing is recommended for semi-finishing external spur and helical gears where the blank configuration permits the clearance for the cutter path. Hobbing is generally the most cost effective means of generating spur and helical gears. Hobbing machines can generate crown and/or taper. Helical gear teeth are generated through a differential gear box or by computer numerical control (CNC)/electronic control.
112.3 Semkfinish tooth geometry The sem’tiinish tooth geometry and accuracy should be controlled by measuring the first part and last part on a qualified measuring machine. Additional parts should be measured when a cutter is changed or the machine setup is changed. Generally, gears are semi-finished by cutting on a hob or shaping machine but may be ground “from solid” (directly from a blank), or milled, forged, rolled, etc.
112.4 Heat treat considerations Consideration should be given to heat treat It may be required to adjust the distortion. measurement over or between pins to accommodate dimensional change in the carburizing and hardening cycle. A lead adjustment may also be required to counteract lead variations introduced by Generally, a pilot lot should be heat treat. processed through the heat treat cycle to confirm the effects of hardening. The larger the gear diameter and the longer the face width, the more critical this process may be. For consideration of microstructure, see clause 9 of this document. Nitriding and induction hardening of some aircraft gears are also used. These processes reduce much of the distortion of the gear teeth and the gear blank experienced during the carburizing and hardening process. The nitriding process produces a thin, very hard, brittle layer in some materials that may cause damage in certain applications. Some manufacturers allow this material to wear off during break-in. Others remove it by either light grinding or by electrochemical machining (ECM). This non-abrasive
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process removes the white layer with chemicals that slowly dissolve the surface. A non-destructive chemical check can be performed to determine if the white layer has been sufficiently removed. 112.5 Gear blank control Gear blank dimensional control before, during, and after heat treat is important. The operation sheets are used to control the manufacturing process, including the dimensions and tolerances for the surfaces used to support the work holding fixtures and the surfaces used to contact the quench die where applicable. The gear tooth accuracy is directly related to the control of the tooling points of the gear blank. In addition, the contact points related to the quench die will dictate the consistency of the gear teeth through the hardening,process. 112.6 Grinding gear teeth. Finish grinding gear teeth is important to provide accurate and consistent gears. Selecting the grinding machine may be the most important decision to assure the gear teeth will meet the engineering drawing requirements for geometry and accuracy. Every gear is unique and one machine may be able to perform better than others. The following are some suggestions in selecting a gear grinding machine: - A shoukier or physical obstruction adjacent to the gearteeth restrictstheselection of a machine type. Form grinding permits the use of small wheels in this case; - Gears which require the root fillet radius to be ground restrict the use of some gear grinding machines; - Most aerospace gears have geometry and accuracy tolerances requiring gear grinding machines to be maintained to excellent condition. lt is not unusual for gear designs to require AGMA Class No. 13 and 14. In order to meet AGMA Class No. 14 successfully, consistently, and economically, it is recommended to use electronic and CNC grinders, or other machines that have demonstrated they can meet this requirement. Temperature control of the grinding machine and the ambient temperature is important to obtain the specified accuracy;
64
- The selection of the grinding wheel and the method of dressing the wheel are factors in controlling the involute form and the surface finish. The wheel abrasive, grii size, hardness, bond, coolant oil, wheel surface feed, and machine feed rates are all factors in the performance. Plated cubic boron nitride (CBN) grinding wheels can be used successfully to grind precision gears with good involute and surface finish. 11.2.7 Measurement and control All precision gears should be measured and controlled based on engineer requirements. Special requirements will include involute, lead, pitch, and index variation and should be measured on a precision certified measuring machine. As the precision and accuracy requirements of the gear increase, so does the accuracy requirements for the measuring system. A calibration method should be established to certify the accuracy of the measuring machine. 11.2.8 Other measurements In addition, the major and minor diameter (whole depth), root fillet radius, and measurement over or between pins and/or balls (tooth thickness) should be measured and controlled. Surface finish, roughness, and waviness are important elements to be measured and controlled. Catburized, hardened, and ground gears should be evaluated for abusive grinding (burning) by surface temper inspection. 112.9 Deburring Deburring and radiusing the edges of gear teeth shoukl be accomplished using care not to damage the working surface of the gear teeth and use caution not to abuse (overheat) the surface, which may cause surface temper on the edges. 11.3 Bevel gears This document will address spiral bevel, straight bevel, and Zero1 bevel gears. It should be understood that the spiral bevel gear tooth form is a three dimensional curve, which is not traceable to any known curve. The geometry of spiral bevel, straight bevel, and Zero1bevel gears is traceable to the machine that is used to manufacture the tooth form. Most precision aerospace bevel gears are ground using modified roll. Wiih the advent of the CNC controlled spiral bevel gear grinding machines, which have the abilii to grind other tooth fans, the aerospace industry will continue to use
AGMA 911-A94
the modified roll system for most spiral bevel gears for some time. Straight bevel gears are usually not ground because there are limited machines available to grind straight bevel gear teeth. Note: The majority of aerospace bevel gears are designed and fabricated using the Gleason system.
11.3.1 The source for data The origin of the machine settings used on the cutting and grinding machines are developed by using the many computer programs available to the industry. Initially, the Engineer will design the gear mesh based on the requirements of the application. The power, velocity, tooth load, shaft angle, ratio, temperature, the mechanical application, and many other characteristics are considered. The following is some of the basic data used by the manufacturing engineer who will develop the machine settings necessary to meet the engineering design: 11.3.2
Number of Teeth; Pressure Angle; Diametral Piich; Spiral Angle; Face Width; Shaft Angle; Whole Depth; Working Depth; Addendum; Root Angle; Root Fillet Radius; Addendum Angle. Develop flank form geometry
Once the basic gear requirement is established by Design Engineering, the Gear Manufacturing Engineer develops the toothflankform geometry by using a computer system. This is accomplished by conducting the following computer analysis: a) First, a dimension sheet which includes the basic dimensions of the bevel gear and pinion and also includes some of the strength and durability factors for the design must be established based on the Engineering requirements; b) Second, a tooth contact analysis is developed. The computer has a working model of the generating motion, including the kinematics of the meshing of the gear and pinion teeth. In principle, this program develops the tooth flank form geometry on the computer in the same mannerasthe machine operatordoes except it is
instant, and the Engineer has a more scientific approach. The Engineer has the opportunity to change any one or all of the eighteen machine settings, which control(s) the gear tooth geometry of the gear, and the eighteen machine settings which control the pinion. Conventionally, the gear is cut and ground spread blade (grinding or cutting both sides of the tooth at the same time), but the Engineer has the option to produce the gear single side in the same manner as the pinion. The tooth contact analysis will display the lines of contact and the path of contact at the toe, heel, and mean of thetooth. In addition, the tooth contact analysis displays the motion curves (transmission error) at the toe, heel, and mean of the tooth. During the development of the tooth contact pattern required by the development of the tooth contact analysis, the Manufacturing Engineer is evaluating the following: i. Lengthwise curvature; ii. Tooth bias; iii. Tooth profile curvature; iv. The effect of the blade angle (longer at the dedendum and shorter at the addendum, etc.); v. Length of the toe versus the length of the heel; vi. The path of contact; vii. The profile width at the toe versus the profile width at the heel. 11.3.3 Engineering skills The development of the tooth contact analysis requires either second or third order changes in order to correct for an undesirable pattern (load distribution). The Manufacturing Engineer’s level of experience and understanding of the gear mesh application is directly related to the level of success in this work Generally, further development work is conducted at the time the gears are generated (cut and ground). 11.3.4 Loaded tooth contact analysis The Manufacturing Engineer has an option to run the loaded tooth contact analysis program. This program is similar to the standard tooth contact analysis but can display the line and path of contact
as well as motion curves based on seven diierent loads. This program may further define required
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AGMA 911-A94
changes to the flank form geometry to improve the load distribution. This would require changes to the input data for the tooth contact analysis.
computer system. It will also provide the machine changes required to correct the errant tooth form.
11.3.5 Summary of machine settings
Manufacturing bevel gears requires a good manufacturing plan. The Manufacturing Engineer should have a good technical background in bevel gear manufacturing. The manufacturing plan should use good logic for processing the gear, taking into consideration the following:
When the tooth contact analysis has been developed to the satisfaction of the gear Manufacturing Engineer, the summary should be produced. This data provides the machine settings for cutting and grinding the gear and pinion. In addition to machine settings, basic gear data is given as well as settings for the Blank Checker, which measures the face and back angle, root depth, and root angle. Proportional changes are given for making second order changes to the pinion. 11.3.6 Grinding sequence Running the Grinding Sequence Program will provide information for sizing the grinding wheel, type of grinding wheel, feeds and speeds, and other sequence data necessary to grind the gear and pinion. 11.3.7 Fillet details The gear Manufacturing Engineer has the option of running “Tooth Profile and Fillet Details” program. This program will define the generated root fillet radius of the gear and pinion based on the generator and grinding machine settings. 11.3.8 Current technology Todays current technology permits defining the working flanks of spiral bevel gears scientifically using a three-dimensional, coordinatesystem. The computer system will define the theoretical tooth form established in the special analysis file located in the mainframe computer, while developing the tooth contact analysis and summary of machine settings. The theoretical coordinate data established on the computer can be downloaded to the coordinate measuring machine for use in measuring theflankform of the spiral bevel gear teeth. With the software system, it is possible to get first and second order corrections to the machine settings used to produce the gear teeth. That is, a gear produced on a cutting or grinding machine can be measured on a coordinate measuring machine using the computer program. It will define the
variance of the machined surface as compared to the theoretical or measured data stored in the
66
11.3.9 Manufacturing bevel gears
a) The dimensional tolerances on the Engineering drawing; b) The gear tooth geometry and accuracy; c) The heat treat distortion and processing; d) The finish processing, including deburring, plating, and final inspection; e) Each operation should be described clearly and concisely using picture sheets when appropriate. 11.3.10 Generating Generating the gear teeth prior to carburizing and hardening is important. The primary considerations are as follows: a) Controlling the flank form geometry to permit minimum stock removal for finish hard grinding the gear teeth; b) Control the root depth and root fillet radius to permit proper stock for finish grinding the gear teeth; c) The tooth accuracy, pitch, and index variation should be controlled to provide the requirements for hard grinding the gear teeth; d) The surface finish should be controlled to assure that the cutter flats and scratches do not endanger cleaning up the tooth surface while finish grinding gear teeth. 115.11 Heat treat distortion Consideration of heat treat distortion during all phases of gear blank fabrication and generating the bevel gear teeth must be maintained. Carburizing and hardening gears, which causes the gear blank to distort, has always proven to be difficult, especially on gears with large diameters and thin cross sections. The spiral angle of spiral bevel gear teeth tends to unwind during heat treat. To control the stock removal of the carburized gear teeth, the
AGMA 911-A94
face and back cone and crown diameter must be maintained within close limits. Quench dies are generally used on large gears by placing contact pressure on the outside diameter, inside diameter, and top face angles during the quench cycle to reduce some of the distortion. Gears must be flat and round after hardening. Nitriding and induction hardening are also used for some bevel gears. These processes reduce much of the distortion of the gear teeth and the gear blank. 11.3.12 Control after hardening Controlling the gear blanks after hardening is accomplished by using a good manufacturing process and tooling. The pitch line of the gear teeth must be concentric to the mounting diameter and mounting surface. This part of the manufacturing process is critical in order to maintain stock removal within limits of the case depth and surface hardness on the gear teeth. The accuracy of the gear teeth are traceable to the accuracy of the gear blank. 11.3.13 Grinding bevel gear teeth Grinding the spiral bevel gear teeth is generally the last machining operation and is among the most important. Using the grinding machine that can meet all the requirements of the engineering drawing is the first and most critical step. Other considerations are as follows: a) Using the established summary and final machine settings for the gear grinder; b) Controlling and recording the machine settings used to grind the gears; c) Stock dividing the gear to the grinding wheel to remove equal stock from the flanks and control stock from thwoot of the gear tooth while maintaining the root depth requirement; d) Use the grinding wheel, which maintains form and produces the required surface finish without tempering the gear tooth. 11.3.14 Measurement of bevel gears 11.3.14.1 Tooth flank form Measuring and controlling ground bevel gear teeth is among the important steps in fabricating bevel gears. The recommended method for measuring
bevel gear teeth flank form is using the coordinate system.
This method is the most scientific and
accurate system. An alternate method is using the rolling method on a test machine for taking tooth patterns. The tooth pattern record should include vertical (V) and horizontal (H) measurements to the toe and heel to control bias, and (V) only measurements to the toe and heel to control the pattern length (see AGMA 390.03a). Additional V&H measurements should be taken to control the tooth profile width. 11.3.14.2 Other measurements The root depth and root fillet radius must be measured and controlled to meet the Engineering drawing requirement. The surface finish, pitch and index variation, and face and back angle of the gear should also be measured and controlled 11.3.47 Deburring and radiusing Deburring and radiusing gear teeth must be accomplished, being cautious not to temper the tooth edges after the gear teeth are ground. Abusive grinding during the debut-ringand radiusing can be detrimental to the strength of the gear tooth. Surface temper inspection should be performed after final machining, including deburring to assure that the gear is free of temper. Automated deburring and brushing are the preferred methods. NOTE - One way to avoid surface temper is to adequately chamfer the gear teeth prior to hardening.
11.3.18 Final processing The final processing of the gears, which includes surface temper inspection, plating, shotpeening, magnetic particle inspection, and coating must be conducted according to approved operating procedures. 11A Stress relief treatment Stress relief should be performed on all gears that have been carburized and hardened as soon as possible after grinding and should be performed prior to any subsequent processing. 9310 steel parts should be stress-relieved at 275 to 300 “F (135 to 150 “C) for at least one hour (depending on the cross section of the gear) followed by still air or oven cooling to room temperature. Stress relief treatment will minimize the possibility of residual stress cracking. Since the possibility of such cracking increases with time between grinding and stress relief, it is recommended that stress relief be performed within eight hours of grinding. 67
AGMA 91-l-A94
12 Gear inspection 12.1 General This clause describes the recommended methods, practices and controls used for the assurance of the desired quality level of aerospace gears. The clause is divided into two parts and discusses spur and helical involute gears and bevel gears. The dimensional and gear tooth element inspection of aerospace gears is the same as for gears used in other applications. However, the controls which maintain the integrity, reliability and life are more stringent and more supportive dccumentation and records are required. 12.2 Spur and helical involute gears The inspection methods and practices for the inspection of spur and helical involute gears are comprehensively discussed in ANSVAGMA 2000-A88, Gear Classification and inspection Handbook, Tolerances and Measuring Methods for Unassembled Spur and Helical Gears (including Metric Equivalents). The following covers additional considerations for the inspection and control of aerospace spur and helical involute gears. 122.1 Identification/traceability Because the integrity and lie of aerospace gears is so important, and in the event of a failure having to
retire other gears from service, it is recommended that all gears be permanently identified with a lot number or with an individual serial number. The lot number or serial number should be traceable to a raw material heat and melt number and be in evidence throughout manufacture and be documented at significant manufacturing, inspection or processing operations. Traceability with serial numbers can significantly assist in the investigation of any problem or failure. 12.29 Heat treat verification The verification of final heat treat results is of the utmost importance, particularly if the gears are processed to produce a high hardness layer such as by carburizing or nitriding. A metallurgical laboratory analysis of an actual part, part section, or coupon (see figure 29 for coupon example) representing the gear teeth and processed with the lot of gears is the recommended method of assuring the results of the heat treating process. This analysis would be traceable to the lot either by permanently ident’@ing the part with the heat treat lot number or by documentation of part serial numbers. The laboratory analysis would also be used to calculate the minimum and maximum amount of stock that can be removed to meet the specified finish depth of the hardened layer.
Surface A, TYP, 2 PL
Drill thru 306 (0.187) diameter
Figure 29 - Heat treat coupon
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122.3 Hardened layer finished depth When required to assure that the finished depth of the hardened layer meets specification, the recommended method is to monitor through process control at the grinder by stockdividing and recording the machine in-feed setting at the first contact with the gear and the final machine setting. The stock removal can then be easily calculated. This method proves the entire process of the gear from the calculated root and flank stock through cutting, heat treat, quench and final gear grind. 129.4 Surface temper inspection Inspection of high hardness ferrous gears, particularly those with carburized surfaces, for abusive grinding should be accomplished by in-process surface temper inspections during the grind operations. It is also recommended that all carburized gears be resurface temper etched after all machining and deburring operations are complete. A discussion of surface temper may be found in AGMA 230.01, Surface Temperlnspection Process. The requirements for this process should be mutually agreed upon by the manufacturer and the purchaser because failure to detect abusive grinding can have serious consequences. 122.5 Inspection flaws
for surface or subsurface
Essential to assurance of the reliability, lie and control of the cost of gears is the need to verify the integrity of their material condition at various stages of manufacture. If there is a concern that the raw material, i.e. forgings or bar, may contain undesirable characteristics such as laps, seams or inclusions they should be non-destructively tested prior to or immediately after final heat treatment. The most commonly used tests are magnetic particle, which detects surface or slightly subsurface defects, or ultrasonic which will detect all defects through the material. If gears are shot peened, a non-destructive test of the shot peened areas prior to peening will assure that no defects open to the surface will be peened closed. An appropriate final non-destructive test should be performed after all manufacturing, plating and coating operations are completed to assure that no
manufacturing, plating or coating operation has induced defects that would be detrimental to the use of the gear. 12.2.6 Inspection of root fillet radii Inspection of the root fillet radii should be accomplished by the use of a magnifying tracer instrument which charts the radius form, or by the use of a cast and optical comparator. To assure maximum tooth strength, the fillet radii should meet specifications for size and smoothness. 12.2.7 Verification peening
of dimensions
after shot
Precision gear configurations that are shot peened should have close tolerance features such as diameters, lead and runout m-inspected after shot peen to assure that no undesirable dimensional change has taken place. Refer to clause 10 for other shot peen requirements. 12.2.8 Conformity inspections Assurance that a first run gear has been manufactured correctly to the design requirements may be accomplished through a comprehensive destructive test of a completed part by experienced laboratory personnel. This activity can also be conducted on a periodic basis through the manufacturer/purchaser agreement to assure process control. 12.3 Bevel gears The inspection methods and practices for the inspection of bevel gears are discussed in AGMA 390.03a, Gear Handbook - Gear Classification, Materials and Measuring Methods for Bevel, Hypoid, Fine Pitch Wormgearing and Racks only as Unassembled Gears. The following covers additional considerations for the inspection and control of aerospace bevel gears. 12.3.1 Master gears To successfully maintain proper pattern position and interchangeability during the production of bevel gears, it is necessary to establish master gears during preproduction development of the gear sets. When the proper working pattern is achieved at least three sets of mating gears should be selected by evaluating the tooth contact patterns on bevel gear test machines and designating them as master gears at the following levels:
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AGMA 91%A94
a) The top level master set is selected as the best set to duplicate successfully tested gears; b) The second level master set is usually designated as the reference master set and is periodically tested with the top level master mate to assure that the correct contact pattern is maintained for transfer to the third level master; c) The third level master sets are used to test the production gears for contact pattern position and backlash and are normally called working masters or working control masters. To preserve the patterns from the master gears for wear comparisons at a later date, tape transfers of the pattern positions are made using a colored compound on the gear teeth during testing and lifting the contact pattern from the tooth with a transparent tape to a white hardboard card.
Wih the proper equipment and software, pitch, index, spacing and the finished tooth surface tooth form, as well as other characteristics, may be very precisely measured. Measurement of the tooth flank form consists of a digitized map of the working surface of either flank that is compared to a computer stored digitized master tooth. This method of tooth flank form measuring is being used successfully in the gear industry to maintain process control during manufacture. Also, the technology is available to allow automatic calculations from the digitized data to corrected machine settings for optimization of tooth form. 12.3.6 Surface temper inspection Refer to 12.2.4. 12.3.7 Inspection flaws
for surface or subsurface
Refer to 12.25.
The top level masters should be well protected, preserved and used only to test second level masters.
12.3.8 Inspection of root fillet radii
Second level masters may require replacement occasionally, dependent upon production volume and the frequency of testing with third level masters.
12.3.9 Verification peening
Refer to 12.2.6. of dimensions
after shot
Refer to 12.2.6.
Third level masters must be controlled closely because they are used frequently to test the working production gearing.
12.3.10 Conformity inspections
12.3.2 Identificatio~nkaceability
13 Rocket and space getiring
Raw Material Control and Traceability - Refer to 12.2.1.
13.1 introduction
Refer to 12.2.8.
Verification of Stock Removal - Refer to 12.2.3. However, the calculation of the stock removal may be somewhat more complex due to the mechanics of the bevel gear grinder.
Gears are in common use in space vehicles of all types. As experience is accumulated in space, it becomes increasingly clear that space imposes its own set of rules for survival. As the distance of operation from the earth increases, the environment becomes a dominant factor in the design equation. The behavior of many materials changes in the vacuum of space, and these changes must be considered in the design of space mechanisms.
12.3.5 Coordinate measuring machine inspections
Vehicles designed for operation in space fall into three general categories:
12.3.3 Heat treat verification Veriication of Heat Treat Results - Refer to 12.2.2. 123.4 Hardened layer finished depth
In addition to the measuring methods and practices for bevel gears that are described in AGMA 390.03a, coordinate measuring multi-axis
a) Rockets - contain single use, short life span, high power gearboxes or mechanisms. These units are expended by a single flight;
machines may be used to inspect many features of
b) Reusable spacecraft and space planes -
bevel gears.
gearing and mechanisms must be suitable for
70
AGMA 911-A94
repeated reuse over a series of flights and launches. Operation in a harsh environment usually involves a short mission duration typically measured in days or weeks; c) Space Station and Satellites - geared systems for these applications must survive their operating environment for extended periods of time (to more than ten years). Functional reliability is paramount and there is no opportunity for preventive maintenance. 13.1 .I Rocket gearboxes One of the first applications of gears for space was in the liquid fueled rocket. The turbopump gearing on these vehicles transmit the power to pump the thousands of gallons of fuel consumed during the first few minutes of flight. All of the considerations for light weight power gearing outlined in clause 4of this information sheet apply to rocket gearing. The power density for rocket gearing is extremely high since the mission time is short and minimum weight is critical. In qualification tests, these gears must function successfully for approximately 5000 seconds at full power. 13.1.2 Spacecraft and space planes These craft embody the use of all the usual mechanisms found on aircraft, such as control surface actuators, bay door actuators, etc. In addition, there are mission specific actuators used in the deployment of satellites and other payloads from the cargo bay. When the mission involves manipulation of large bodies, the vehicle, such as a space shuttle, is fitted with a robotic arm. Mechanisms aboard space shuttles and planes are typically designed for repeated usage, and therefore must withstand the vibration and shock of repeated lift-offs and landings in addition to the function they perform on a space mission. Since a shuttle operates at an attitude less than 150 miles above the earth’s surface, the environment includes exposure to highly reactive atomic oxygen and micro-meteorites. 13.1.3 Space station and satellites Although geared systems in this last category of
mission is extended and can be in excess of 10 years. As a consequence, the opportunity for inspection and refurbishment does not exist, and the mechanisms will be subject to degradation of lubricant from evaporation and erosion, and extended wear on working surfaces. 13.1.4 Gear forms for space use Many gear forms have found a place in space mechanisms. These include extensive use of spur gears, some bevel, helical and worm gearing, and an occasional proprietary tooth form. Since light weight structure is of paramount importance, gear forms requiring high precision mountings are more diiicult to apply successfully in the space mechanism. Similarly, forms that are relatively heavy for the torque transmitted, or forms that exhibit a high level of sliding at their contacting surfaces, are less suitable for the space mechanism. Motors which power space mechanisms are themselves small and usually have only modest torque outputs. As a consequence, the efficiency of the drive is of prime importance. In some drives, such as those used on robotic arms, provision for backdriiing in case of a collision or momentary overload is a requirement. To achieve backdriieablilty in a gear train whose transmission ratio may be in excess of 18OO:l points to the use of high precision spur or helical gears. Althoughoperating ina weightlessenvironment, the payload mass which must be manipulated is often large, and output torques of a few thousand foot pounds are not uncommon. The contact stress capacity of space qualified lubricants imposes serious constraints on usable tooth loads. Hence in high reduction dries, the designer must consider dividing the tooth load over a number of gear meshes. The planetary gearbox is often utilized to meet these demands. 13.2 Lubrication The approach to lubricant and material selection for spacecraft gears must be tempered by consideration of the environment in which they will be expected to operate. Conventional oil and grease lubricants are not suitable for space application since in the presence
space vehicle may not be especially different than
of a high vacuum, normal lubricants will outgas and
the ones found on a shuttle, the duration of the
their usual lubricating properties will be destroyed.
71
AGMA 911-A94
Some dry compounds, such as graphite, which exhibit lubricity at sea level become extremely abrasive in a vacuum. Other lubricants, although technically usable in a vacuum, will migrate and can contaminate sensitive optical and electronic equipment. Clean, non-lubricated metal, under the influence of contact pressure and sliding in a vacuum, can cold weld. Subsequent separation of welded surfaces will tear the welded spots apart leaving severe craters and pits. Fortunately, certain grease and dry compounds have been identified, which offer a solution to the lubrication problem and have been used with reasonable success. These lubricants fall into two categories: dry and wet. Each has its own limitations. 13.2.1 Application To be effective, a lubricant must remain in the load area in order to prevent metal to metal contact on working surfaces. In ball bearings, the churning action created by the rolling balls tends to redistribute lubricant on a continuous basis. In gears, the tooth action tends to wipe or squeeze the contact area free of lubricant, especially in slow, high torque meshes. A planetary gear train will minimize such lubricant loss since the internal gear will serve to contain the lubricant, and the circulating planets will redistribute some lubricant among working meshes. A successful, long life, wet lubricated spur gear mesh in space remains to be demonstrated. 13.2.2 Dry lubricants One dry lubricant which has been in extensive use is a molybdenum disulphide compound combined with a carrier. The material is applied to the surface to be lubricated, and may be subsequently baked, and finally burnished. The thickness of material after burnishing is typically in the range of 0.0002 inch (0.005 mm). Since metallic geartooth surfaces must be separated by a thin residual film of dry lubricant, the Hertzian contact stress which the lubricant can carry establishes the load limits for the gear train. Once the lubricant film is destroyed, gear failure from scuffing, pitting, and welding can be expected. As a rule of thumb, gear contact stress up to 100 000 lb/in2 (690 N/mrr?) has given good service for low speed, robotic gearing. On meshes
where the pitch line speed may approach 1000 ft/min (5 mls), limiting the contact stress to 35 OOO-
72
40 000 lb/in2 (240-280 N/mrr?) is prudent. Each application should be verified by test in vacuum conditions before being approved for flight. Some typical solid-film lubricants with space history are listed in table 17. 13.2.3 Wet lubricants Both gears and bearings have been successfully lubricated in space with oil and/or grease for short periods of time. However, the behavior .of these lubricants in the space environment imposes some constraints on the design which must receive careful attention. 13.2.3.1 Vapor pressure The vapor pressure of fluid lubricants for space use is extremely important. Space lubricants typically have a vapor pressure of I Od to 10-l” torr (1.35 x 1o-4to 1.35 x 10” Pa) at 1OO°F(38%). Evaporation and migration of fluid lubricants necessitate the use of an extensive sealing system which can be complex and occupy limited weight and space. The more viscous oils typically have lower evaporation rates. 13.2.3.2 Surface migration (creep) Oils vary in their migration characteristics. Mineral oils like Apiezon C have the lowest migration rates, while silicone and fluropolymer oils like Versilube F-50 and Ktytox have the highest. Lubricant creep may be minimized by the use of flurochemical barrier films. However, these chemicals must not enter the load contact area. 13.2.3.3 Greases Greases are oils that have been thickened by the addition of either soaps (e.g. stearates) or fillers (e.g. finely divided silica). In most cases, the thickening agent is not a good lubricant itself, and most of the important properties of the grease are derived from the oil in it. Under vacuum conditions, the oil will eventually outgas, leaving no lubricant and sometimes an abrasive thickening agent. A partial listing of space qualified wet lubricants is shown in table 18. 13.2.3.4 Rocket lubricants Rocket gear trains
have
been successfully
lubricatedwith either oil or engine fuel, as listed in table 19.
AGMA 91%A94
Table 17 - Candidate solid-film
lubricants for space application
Lubricant*
Method of application
Lubeco 905
Coated Sprayed
MoS,, PbS & graphite, + inorganic binder.
Good load carrying capacity. Low coef. of friction. Oil contaminates deteriorate lube.
Viirolube 1220
Coated Sprayed
MoS,, Ag & graphite, + ceramic binder.
Ideal for hi temp applications. Good load capacity. Approx 1000 “F cure temp. limits substrate hardness. Ceramic binder.
Microseal
Impinged on Surface
MoS,, + ceramic binder.
200-I
Moderate load capacity. Insensitive to oil contamination.
Electrofilm 2396
Sprayed/ Coated
MoS,, & graphite with sodium silicate binder.
Used on moderately stressed components. Sensitive to oil contamination.
Electrofilm 4396
Sprayed/ Coated
MoS, & graphite, phenolic binder.
Good load carrying capacity.
MoS2
Sputtered
MoS, ion bombardment in vacuum.
Good. solid-film coating. Fair to good load capacity.
Molykote
- -
Sodium Silicate binder, MoS, and graphite.
Gear and bearings, deployment mechanisms.
Duroid 5813
- -
Teflon reinforced with glass MoS,
Self-lubricating material used as thrust washers, bearing element separator, etc.
Vespel
--
Polymide
Used as thrust washers, bearing element separator, etc.
Lead
Ion plated
Lead
Good load capacity.
x-1 5
Constituents
Remarks
* This information is given for the convenience of users of this AGMA information sheet and does not constitute an endorsement by AGMA of the products named. Equivalent products may be used if they can be shown to lead to the same results.
13.3 Gear materials for space application In selecting a gear material, the intended duty, lubricating media, service life, and environment are all factors to be considered. 13.3.1 Rocket gearing The high speed, high power gearing found in turbopump drives requires more premium grade material than is used in helicopter and turboprop units and the suggestions of clause 4 apply. 13.3.2 Spacecraft gearing Some of the dry film lubricants applied to spacecraft
gearing are hygroscopic and may therefore promote rusting of the substrate material. To avoid rust contamination, corrosion resistant steels have found wide application in steel spacecraft gearing and structures. For high loads the material can be gas plasma or ion nitrided. On some designs precipitation hardening or maraging steels heat treated to the 44 - 48 HRC hardness range are used. In general, tooth wear rate is found to be proportional to the square of the Hertzian stress. One such material which has found extensive use for space gears is AMS 5617 in the precipitation hardened condition.
73
AGMA 911-A94
Table 18 - Candidate fluid lubricants for space application
Lubricant*
Viscous VI, Vat-Kote :Hydrocarbon base oil)
T
T
Viscosity CSt
1OO°F 210°F 138%) (100°C) 100
10
Space history and remarks
Other properties
Vapor pressure 6 x 1O-” torr (8 x 10 -‘Pa) at
77°F (25 “C).
Used on OS0 IIINTELSAT III, IV for bearings, slip ring assemblies, DC motor brushes, etc.
Versilube F-50 Vlethylchloropheny ISilicone oil
52
16
Pour point -100°F (-75 “C). Useful operating temp. in air is -1 OOOFto +450°F (-75 to +230°C).
Used on Gemini, Manner 3 & 4, OGO, Nimbus, Ranger, Mercury & others. Used on ball bearings & gears.
Aeroshell7 Soap thickened diester
--
3.1
Pour point -100°F (-75. “C). Useful operating temp. -100 to +3OO”F
Used on Ranger 6 - 9.
(-75%
to +15O”C).
Apiezon “c’ 100 Molecularly distilled hydrocarbon oil.
100
10
Pour point +I 5°F (-9°C) Vapor pressure 1O* ton (1.3 x 1OGPa) at 68OF (20 “C).
OS0
Braycote 601 Grease
140
45
Dropping point 1500°F (820 “C). Vapor pressure 1 0-gtorr (1.3 x 10v7 Pa) at 38.8”F (3.8 “C).
Perfluorinated polyether with tetrafluoroethylene telomere
This information is given for the convenience of users of this AGMA information sheet and does not constitute an endorsement by AGMA of the products named. Equivalent products may be used if they can be shown to lead to the same results. l
Table 19 - Working fluid lubricants For lightly loaded applications, plastic gears have been applied to space gearing. These are used in combination with stainless steel gears. Plastic gears giving the lowest wear rate have been MoS, filled polyimide. Hard anodized aluminum gears have been used on a few satelliies, however the coating is subject to cracking under high loads. Such gears have severe limitations when run against the same material. Improved performance will result if the hard
anodized aluminum gear is run against a plastic gear in the presence of MoS,.
74
Lubricant MIL-L-7808 MIL-L-23699
1 ~Space history Aerojet ~ Titan Rocketdyne Turbopump Mark 3
Fuel additive ~ Kerosene & 3% Zinc diacoele ~ Delta engine RS-27 MIPHI Phosphate i Gaseous hydrogen Liquid oxygen
RL-10 Pratt & Whitney I Aerojet
AGMA 911-A94
Annex A (Informative) Spur gear geometry factor including internal meshes [The foreword, footnotes, and annexes are provided for informational purposes only, and should not be construed as a part of AGMA 911-A94, Design Guidelines forAerospace Gearing.]
A.1 Introduction
b
is dedendum of gear, in (mm);
Many texts dealing with the capacity of spur gears neglect a discussion of internal gear meshes. Internal gears are especially important in aerospace applications since planetary reduction stages are widely used. The definitions of the involute portion of the internal gear are the same as the imaginary external gear comprised of the tooth spaces of the internal gear. The equations presented herein are written in a form to be compatible with a computer program.
%ate
is addendum of mate, in (mm);
*P
is circular tooth thickness of gear, in (mm);
To determine the geometry factor, J, for bending strength of an internal mesh, two general cases must be considered as follows: - External gear with internal mate: - Internal gear with external mate. The equations presented below are written in a general form and include “signed” integers, & and AM, for the gear and mate respectively. The signed integers used in the J factor equation permit their use for both internal and external gears. & and AM are defined as follows: = +I for external gear; & = -1 for internal gear; 4 AM = +1 for external mate; AM = -1 for internal mate. The “gear”is defined, as the member for which the J factor is desired, whereas the “mate” is defined as the mating gear. All calculations are based on a diametral pitch of 1 .OO and a standard center distance. The equations apply to a true radius ground fillet (not generated).
NOTE - tp may be any value so long as:
BLm~
+ tp + *pnme =n; is fillet radius of gear, in (mm);
‘f
is circular tooth thickness of mate, in (mm); is total backlash, in (mm). %otd A.2 Geometry factor calculation *pmate
R
NG /2 = pitch radius of gear, in (mm);
Rr
R-(k) b = root radius of gear, in (mm);
Rb
R cos Cp= base circle radius of gear, in (mm);
Rmate
N naate/2= pitch radius of mate, in (mm);
RGmate
R mate + (AM) amate = outside radius of mate, in (mm);
Rbmate
Rmate cos q5= base circle radius of mate, in (mm);
pb
7ccos q5= base pitch, teeth/in (teeth/mm);
c
(AM)R + (4) Rmate = center distance, in (mm);
tan+*
=
(&#$, --
@d(Ad~R~
Rb
The following basic data is required prior to starting the calculation.
+ h%f~ c SWJ
is pressure angle:
is number of teeth in gear; NG Nmate is number of teeth in mate;
. ..(A.l)
&
NO?E - all angles are in radians.
9
&
where tan @h
is tangent of pressure angle at highest point of single tooth contact on involute.
75
AGMA 911-A94
or
+(AG) wq -(AG) - *i
*d % =fan&
Al=
R;+ [d[R,
+&I
lj-12-Rb+k)
p12 . ..(A.7)
mb ...(A.4)
for internal or external gears with tangency point above the base circle, and
tt, is tooth thickness on base circle (may be
cos lY,=$ I A, for all internal gears and for external gears where the tangency point is above the base circle diameter
Rx =*
n
where negative for internal gears whereAo =-l), in inches (millimeters); rjn is angle which the normal force makes with a line perpendicular to the tooth centerline at highest point of single tooth contact, in radians;
where r, is the pressureangle at the intersection of the fillet and involute; A, is the radius to tangency point of fillet and gear tooth profile, in inches (mm).
Rx is radius on tooth centerline to point of application of worst load, in inches (millimeters).
To determine coordinates for the center of the fillet radius:
An undesirable condition is a design where the fillet radius for an external gear is tangent to a radial line below the base circle. This can happen for small fillet radii, small pressure angles, and small numbers of teeth. Since the tangency point of the fillet is below the base circle, there can be no involute from the end of the fillet to the base circle. The equations are valid for this condition if a “radial line”connector is assumed between the base circle and end of fillet. For internal gears, it is obvious that this condition can not occur since the fillet must always be at a larger radius than the base circle. For an external gear, & = +l and the check for this condition is as follows: R; -R2 (for external gears only) rf ’ 2R rr ...(A.!$ If the above expression is true for an external gear, then the tangency point of the fillet is belowthe base circle diameter and a radial line connector is assumed. To calculateA, for this case, equation A.6 is used for external gears only. For all internal gears, or for external gears whose fillet is above the base circle, the above check is false (for external gears) and equation A.7 is used to calculate A,. A, = y/m
@=&%;I ...(A.6)
for tangency point below base circle diameter, and
cos r, = 1.0000 where the tangency point is below the base circle diameter
76
1 aI= invI;+
l+bk) 2
1 ‘b 2Rb
x’=A,$n
a’
Y’=A,cos
a’
-6%)
*b’ . ..(A.8)
A =a’+I’, aa= x’ -64~)
7 cos A
. ..(A.9)
bb=y’+&)
rf sinA
. ..(A.8)
where aa isxcoordinate of fillet radius center, in inches (millimeters); bb is y coordinate of fillet radius center, in inches (millimeters). Figure A.1 illustrates some of the nomenclature used for external spur gears while figures A.2 and A.3 are for internal gears. Figure A.4 demonstrates the special case where the fillet center is below the base circle diameter and the fillet and involute are connected by a radial line portion. Afterthe coordinatesof the center of the fillet radius have been calculated, the next step is to determine h and t, the height and thickness respectivety of the tooth at the critical section. Figure A.5 illustrates h and tfor external and internal gears.
,
AGMA 911-A94
& space
bb
IB B--Jc Nl%’ -L
--I
r-
‘. -
I
Figure A.1 - External spur gear nomenclature
77
AGMA 911-A94
&tooth A!--tI
Figure A2 -Angles
78
1II T i A
locating fillet radius center for internal gear
AGMA 911-A94
k space
ctooth
bb
I I
++‘i
4\ \
1RX I I
I Base
circle I
&space
jFx*Y
1
1
T&D
\\II I\ 1 1IIL ri cp. I
I
l
I
I
J
I
I 1I
Figure A.3 - Internal gear nomenclature
79
AGMA 911-A94
I I I -1 a ---+jiji
fillet
Internal gear . SPACE
Figure A.4 - External spur gear with fillet tangency point below base circle The equations for the calculation of h and t are different in the zone of the involute and in the zone of the fillet and are given below for the generalized radius, A. In the zone of the involute: A is radius to the point on involute, in inches (millimeters);
External gear
A lies between A1 and Rx for zone of involute. [(AG>O)AND(A>AI)AND(A>R~)I OR [(AG < 0) AND (A
cosr= -Rb
...(A.ll)
A
A is radius to point on fillet, in inches (millimeters). [
a =inv r+
1 + (AG) 2
1 ‘b 2Rb
In the zone of the fillet:
80
Figure A.5 - Definition of built-in section of tooth
-(AG) fb’ ...(A.12)
[(&>O)~tA
[(AG< 0) AND (A >A1 11
A lies between Al and R, for zone of fillet.
AGMA 911-A94
[ ~;!+?3;$] u=arctan~+(&)arccos ...(A.l3) NOTE-if (A1sAsRb )thena=a’ (radiallineportion)
h=(&)R,-(&)Acos[(
‘+;AG’>” NG- wo;] ...(A.l4)
t=2Asin
I[(
l+@G z;n 2 9
(&) a]
...(A.%)
Similar equations may be developed forthe case of a generated fillet. After defining the general equations to calculate h and t at any point, it remains to determine the location where the stress will be the highest. The classical method which dates from the Lewis equation was to assume that the tooth had an inscribed parabola with the apex at the load point and which was target to the tooth as shown in figure A.6.
ences other than pure bending which are functions of h and t: - The axial component of the normal tooth load which produces a compressive stress that subtracts from the bending; - The stress concentration factor which is a function of h and t. Therefore, the point at which hand tact is defined as the point which produces the highest total stress and not the highest bending stress. Since all the influencing factors are contained in the geometry factor, J, the object is to determine h and t where J is minimum. J is defined by: 1
J=
cos 4n 1.5 Kt---- cost [ x
1
tan 4% t
. ..(A.l6)
t2
X=4h Kt =H + [+]”
. ..(A.l8)
[$]M
,
where H, M, and L are given in table A.1 . Table A.1 - H, L, and M for use in Kt (from linear extrapolation of Dolan and Broghamer method)
Figure A.6 - Constant stress parabola
Pressure angle $
H
14.5* 20* 22.5 25 30
0.22 0.18 0.1618 0.1437 0.1073
M
L 0.20 0.15 0.1272 0.1045 0.0590
0.40 0.45 0.4728 0.4955 0.5409
* Original Dolan and Broghamer values
Since a parabola has the property of constant bending stress, section B-B, which is at a point of tangency between the parabola and tooth profile will have the maximum bending stress. Note at any other section such as A-A, the stress would be the same as at section B-B only if the tooth were of thicknesst~. Howeverintheactualtoothatsection A-A has more metal outside of tM and therefore has lower stresses. This classical assumption is
Assuming that H, M, and L are straight line functions of the pressure angle, they may be expressed as follows:
based on constant bending stress. However,in the
procedure is used utilizing the equations for h and t
modern definition of tooth stress, there are influ-
which have been previously defined as a function of
H = 0.3255- 0.4167Cp L = 0.3318 - 0.52094 M = 0.2682+ 0.5209$I where $ is in radians. To find the point where J is a minimum, an iterative
81
the general radius, A. A simple method is to divide the section between R, and R, into fBed number of intervals, n , and to start at some point such as R, . J is then successively calculated using a new value of A each time defined by: . ..(A.19)
Figures A.7 through A.9 are for external gears with internal mates for gears of varying pressure angles, while figures A.10 through A.12 are for internal gears with external mates. These curves may be used for preliminary design purposes prior to the detailed Jfactor calculations using the actual tooth dimensions.
J will continue to diminish until a minimum is found, at which point the next value of J will increase. The accuracy of the answer will depend on the magnitude of the interval A. With the advent of high speed computers, the calculation of Jby the above method is easily accomplished. Figures A.7 thru A.12 are plots of the geometry factors for internal/external gear meshes for gears with standard proportions.
Figures A.7 through A.12 have been developed using the equations in thisannex, with the input data shown in tables A2 through A.7. Note that all values shown are for diametral pitch = 1 .OO. Maximum value a = 1 .O was used for the addendum and Br,=O.O4 for the the backlash; minimum value d = 1.24 was used for the dedendum and t = 1.5508 for the tooth thickness.
A=A+(&)A
82
AGMA 911-A94
.60
56
‘-15
20
80 40 60 30 Number of teeth in external gear
100
150
Figure A.7 - Geometry factor for standard proportions, external gears with internal mates, addendum = 1 .OO,dedendum = 1.24, tooth thickness = 1.5508, backlash = 0.04, (I = 20” Table A.2 - Fillet radius used to generate J factor curves in figure A.7 Fillet radius, inches Number of external teeth 17 20 25 35 50 75 100
50 0.613 -------
70 0.613 0.596 0.561 -----
Number 90 0.616 0.596 0.565 0.502 ----
of internal teeth 110 150 0.617 0.617 0.596 0.596 0.560 0.562 0.526 0.512 0.467 0.448 -0.412 ---
200 0.616 0.596 0.560 0.532 0.479 0.426 0.396
300 0.614 0.596 0.560 0.531 0.492 0.441 0.413
AGMA Qil-A94
-60
56
Number of teeth in
.52
.36 ’ 15
20
30 40 60 80 Number of teeth in external gear
100
150
Figure A.8 - Geometry factor for standard proportions, external gears with internal mates, addendum = 1 .OO,dedendum = 1.24, tooth thickness = 1S508, backlash = 0.04, $ = 22.5” Table A.3 - Fillet radius used to generate J factor curves in figure A.8 Number of internal teeth 17 20 25 35 50 75
100
84
Fillet radius, inches Number of internal teeth 110 150 50 70 90 0.546 0.513 -----
--
0.552 0.513 0.494 --mm
0.554 0.532 0.500 0.462 ---
0.552 0.532 0.498 0.463 0.431 --
--
--
--
0.550 0.530 0.499 0.470 0.442 0.415
--
200
300
0.546 0.525 0.498 0.463 0.441 0.419 0.411
0.544 0.522 0.493 0.466 0.444 0.429 0.417
AGMA 911-A94
.60
.56
‘T % t$ -52 al5 -cd -UEs G&8
30 40 60 80 Number of teeth in external gear
100
150
Figure A.9 - Geometry factor for standard proportions, external gears with internal mates, addendum = 1 JO, dedendum = 1.24, tooth thickness = 1.5508, backlash = 0.04, + = 25O Table A.4 - Fillet radius used to generate J factor curves in figure A.9 Fillet radius, inches Number of external teeth 17 20 25 35 50 75
100
Number of internal teeth 90 110 150
50
70
0.474 0.450 -----
0.473 0.452 0.420 i---
0.474 0.454 0.423 0.386 ---
0.477 0.452 0.419 0.393 0.365 --
--
--
--
--
200
300
0.474 0.447 0.421 0.394 0.370 0.341
0.471 0.441 0.417 0.389 0.366 0.346
0.467 0.439 0.412 0.384 0.359 0.337
-a
0.343
0.330
85
AGMA 911-A94
.62
.60
58
8L 5 .56 -is? SE =g.5454
50 .50
.48 ' 40
60
80 100 150 200 Number of teeth in internal gear
300
400
Figure A.10 - Geometry factor for standard proportions, internal gears with external mates, addendum = 1 .OO,dedendum q 1.24, tooth thickness = 1.5508, backlash = 0.04,$ = 20” Table A.5 - Fillet radius used to generate J factor curves in figure A.10 Fillet radius, inches Number of internal teeth 50
70 90 110 150 200
86
Number of external teeth
17
20
25
35
50
75
100
0.412 0.422
_0.425
_0.424
0.431
0.433
0.434
0.437
0.438
0.439
0.444 0.451
0.444 0.451
0.446 0.450
--0.437 0.439 0.448
0.455
0.455
0.454
---0.438 0.448 0.452 0.454
-_ ---0.448 0.453 0.455
__------
0.450
0.453
AGMA 91%A94
I I
.62
I I
Number of teeth in matinn external near ’
75
50
-
.
35 \ .60
17 20 \ \
$
h
\
\
\
F \
\
.58
.56
.52 ~~
.50 40
60
80 100 150 200 Number of teeth in internal gear
300
400
Figure A.11 - Geometry factor for standard proportions, internal gears with external mates, addendum = 1 .OO,dedendum = 1.24, tooth thickness = 1 S506, backlash = 0.04, $ = 22.5” Table A.6 - Fillet radius used to generate J factor curves in figure A.11
I
Fillet radius. inches Number of internal teeth 50 70 90 110 150 200 300
of external teeth l--TnsY 0.333 0.351 0.365 0.368 0.372 0.375 0.374
0.360 0.345 0.363 0.362 0.372 0.372 0.383
yF ---0.344 i 0.357 0.357 -0.366 0.366 0.364 0.372 0.374 0.370 0.383 0.378 0.378 0.374 0.378 0.383
75
100
--
---
0.370 0.375 0.385
0.385 0.382
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AGMA 911-A94
3 5 .62 aao -s
E6 i2g = g.60
58 I
I
II
80
100
I
I
I
150
200
300
400
Number of teeth in internal gear
Figure A.12 - Geometry factor for standard proportions, internal gears with external mates, addendum = 1.99, dedendum = 1.24, tooth thickness = 1.5599, backlash = 0.94, $ = 25” Table A.7 - Fillet radius used to generate J factor curves in figure A.12
I Number of internal teeth
88
Fillet radius, inches Number of external teeth 17 25 35 50 20 ---0241 0241 0260 0262 0.267 --0.270 0.271 0.274 0.280 -0278 0.279 0.276 0.282 0.282 0288 0288 0.288 0.291 0.288 0295 0294 0.294 0.305 0.303 0295 0.298 0.295 0.294 0.306
I
75 ----
-- I 0.288 0.290 0.280
---
AGMA 911-A94
Annex B (Informative) Gearbox Test and Mission Requirements [The foreword, footnotes, and annexes are provided for informational purposes only, and should not be construed as a part of AGMA 91 l-A94, Design Guidelines for Aerospace Gearing.] For gearbox design and test purposes values of B.l Introduction power and speed are needed for each phase of Mission profiles are established to determine load, flight. Figures B.2 and B.3 show typical power and speed, operating time, and environmental requirespeed values versus time for both a fixed wing and a ments for gearbox operation during service. The rotary wing application. designer uses these parameters to size the gearbox while testing is performed to assess the integrity of 10 J_Aad-l ---the design and to uncover any unmanifested F-r --! problems. Gearbox testing is performed in test rigs where possible, and finally in the actual application for certification.
B.2 Mission Requirements Missionsvary depending on the customer, military or commercial, and on the type of aircraft/ engine such as propeller driven, helicopter and turbojet engines. The basic phases of flight are shown, for a commercial aircraft, in figure B.l which displays altitude versus time. The definition for each numbered phase is listed as follows:
_---m-m-------
-
5 20
_
-
-
-
-
-4’o-
-
-
-
-
-
. .
-
-
-i
4
lib
I OO
--
I 20
I 40
I 60
I 80
i 100
% lime
Figure B.2 - Commercial application, turbofan aircraft
1 - Takeoff 4 - Altitude change 7 - Landing 8 - Ground 2 - Climb 5 - Descent Operations 3 - Cruise 6 - Approach Other maneuvers include Holding Patterns while above airports and Go-Arounds during landing attempts. For some military aircraft Loitering is a phase that may consume 25% or more of the mission time.
0 0
I 20
I 40
I 60
I 80
100
% Time
Figure B.3 - Military application, helicopter
Time
Figure B.l - Commercial application, phases of flight
For actuator gearboxes, the highest loads are usually experienced at the lowest speeds and in many cases these loads occur at zero speeds (zero “power”transmitted). These loads are referred to as “holding loads”, “stall ioads”, or “limit loads”. They are def&f by the customer for a specific applica-
89
AGMA 911-A94
tion, and hence, they become a major criteria governing the design. Commercial applications are often more stringent than military in terms of gearbox loading and hours of use. Military aircraft, however, operate in a more severe environment which may shorten gearbox fife for reasons other than load. Figure B.4 represents a 150 hour certification test for a helicopter engine that containsa Power Reduction Gearbox that must be certified also. The test is designed to be more severe in terms of load than an equivalent number of hours of field operation. For example, if power levels below 90 % do not have a significant effect on the accumulated fatigue damage of the gearing, then only 2-l/2 % of the mission time shown in figure B.3 would be applicable to reducing fatigue life, a matter of l-1/2 minutes. By comparison, 50 % of the time, 4 500 minutes, is spent at or above 90 % power in the 150 hour qualification test. This illustrates that sufficient operational hours can be accumulated during a certification test to reveal long term field related operational problems. Figure 8.5 shows power usage spectrumsfortypical small and large military helicopters. In general, as the size of the helicopter increases, the percent time near the top power levels increase.
B.3 Testing Testing is performed to assess the integrity of the gearbox design and is a requirement for aircraft and
engine flight certification. The endurance and environmental testing outlined in the following sections is extracted from the Army Aviation Specification AV-E-9593D, 1934 for turboprop and turboshaft engines. Numerical values are listed to show the stringency of the test requirements which, when completed successfully, will lead to safer operation of the aircraft. Limit and ultimate load testing are included to cover actuator testing.
g60
2
t
$40 I
2:: 0
20
40
80
100
Figure 8.5 - Milita~~~l,icopters, large and 8.3.1 Types of Testing 8.3.1 .l Development Testing Development testing is required to assess the capability of a transmission design prior to qualification or flight. Certain tests, as outlined below, are
Repeat the 5-hour cycle 25 times
Figure B.4 - 150 Hour engine certification test 90
60
% Time
AGMA 911-A94
required to establish and develop basic operational capabilities such as lubrication system tests and gear pattern development tests. Other development tests are conducted to prove out structural integrity, search out “weak links” and gain confidence in the design. The ultimate goal is to verify that the design requirements for reliability and maintainability have been achieved by the time the transmissions are put into service. - Lubrication Systems Tests - Testing at no load and varying speeds up to overspeed conditions. Bearing temperatures are monitored for over-temperature conditions. Oil quantity, pump flow rate and pressure are established. Then power is applied for development of cooling system requirements; - Gear Pattern Development Tests - Gears are tested to determine load distribution at various power conditions. Visual techniques as well as strain gaging are used to determine load diitribution as a function of applied load. Machine grinding corrections are used to assure full contact at the desired power level and to correct for deflections. After testing, final machine corrections are added to the production manufacturing procedure; - Over-torque Development Tests - Tests are conducted at over-torque conditions using accelerated mission profiles to substantiate a higher equivalent operating life and to demonstrate non-catastrophic and fail-safe modes through the use of different inspection and detection techniques. Over-torque testing often accelerates failure modes. This is beneficial in a development test program so that design or manufacturing”changes can be made in a timely manner; - Bench Tests - Used to substantiate systems which have loads induced from external sources, i.e., a shaft and housing which support a propeller or a rotor; - Vibration Surveys - Vibration surveys are conducted to determine resonances in the operating range. Mode shapes for the resonant frequencies are found by holographic interferometry, rap testing with accelerometers, or other methods.
B.3.1.2 Endurance Testing Endurance tests are performed to assess the integrity of the gearbox design and are usually performed with maximum parameter conditions acting simultaneously. - Accessory Gearbox Drive Test - 300 hour endurance/ 600 starts. Simultaneous operation of all drives with each drive subjected to the maximum permissible torque or power rating including the maximum overhung moment and misalignment angle of each accessory. A vibratory and resonant search from engine idle up to 125 % of maximum speed under varying loads. Tests to be conducted at maximum oil inlet temperature and minimum oil flow. The same gearbox is to be used for the Static Test; - Accessory Gearbox Static Test 150% maximum static torque on all accessory pads simultaneously and with the starter loaded to 250 % of max. starting torque for five seconds; - Engine Output Shaft Drive Spline - Demonstrate ability to absorb thrust In either direction equal to 20% of the circumferential force acting on the output drive shaft spline at the maximum continuous torque rating. Perform endurance testing at maximum allowable misalignment and at a speed of 1.15 times the maximum continuous rating. Demonstrate a continuous operating capability of 1.2 times the maximum continuous torque rating; - Oil Interruption Test - Operate engine for 30 minutes at intermediate power (a rating somewhat higher than maximum continuous power) with air only at the oil pump inlet. Then with oil supplied, operate the engine for 30 more minutes. There shall be no damage to any components as a result of the test; - Windmilling Test (backdriving) - 8 hours continuous operation without component damage or excessive loss of lubricant: - Low Cycle Fatigue - A test cycle from engine idle to maximum power and back to idle lasting approximately 5 minutes with a shutdown every 15 minutes. Repeat cycle for a total of 3 750 cycles. The minimum LCF life of the components shall be 15 000 cycles.
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AGMA 911-A94
B.3.1.3 Environmental Testing A number of environmental requirements exist for the safe operation of airborne gearboxes. Some of the Military requirements and reason for testing, as outlined in MlL-STD-31 OC, are listed as follows: NOTE - Numericalvalues are not listed here as they are extensivelycoveredin the MIL-STD.
- Low Pressure (altitude) - Gearbox venting, oil leakage and effect on heat convection and conduction characteristics; - High Temperature - High temperature storage or service conditions; - Low Temperature - Loss of resiliency of gasket materials, congealing of lubricants and tight meshing of gears; - Solar Radiation - Damaging effects on natural rubber and plastics; - Attitude -. Insure proper oil scavenging for all flight maneuvers and for leakage during transportation; - Rain, Humidity and Fungus - Component damage resulting from wet and warm atmosphere; - Salt and Sand - Component corrosionlerosion; - Flame - Fires during storage or operation; -Vibration and Noise- Performance degradation or component breakage and airport operating requirements; - Shockor Impact-Operation and transportation requirements; - Combined Effects - Where multiple conditions may exist simultaneously i.e. Altiiude - combining low pressure with high temperature. Space combining low pressure and temperature with radiation.
induced dynamics. Testing duration is usually from 50 to 200 hours. Drive shafting, cooling systems, tail and intermediate gearboxes, and accessory drives are also tested. After ground testing has shown that no major development issues exist, the transmission is tested in flight. Flight and ground testing continue simultaneously until qualification is complete. The objective ’ of the flight test program is generally fourfold: 1) Structural substantiation; 2) Demonstration of handling qualities and the automatic flight control system; 3) Verification of the propulsion/drive system; 4) Confirmation of the mission equipment. Structural substantiation testing establishes and expands flight envelopes including evaluation of all attitudes and conditions of flight. This includes extreme maneuvers and maximum and minimum loads on all major components. Handling qualities confirms basic data for the controls and control system characteristics and evaluates control reliability and fidelity. Drive system testing confirms structural integrity and measures flight loads required to establish component life. Mission equipment testing includes instrumentation, navigation and communications systems quality, human factors criteria, and related systems capabilities and performance. B.3.1.5 Actuator qualification testing
8.3.1.4 Flight Testing
Testing requirements are determined largely by the part reliability requirements. Test programs often include, but are not limited to the following: - Baseline performance testing; - Thermal vacuum testing; - Vibration/shock testing; - Humidity testing; - Lie testing; - Structural load testing.
Prior to conducting flight testing, development testing has usually been completed or is well along in schedule. Some qualification agencies require tie down testing of a helicopter with the complete dynamic system installed prior to flight test. The helicopter is tied to the ground and the rotors operated to full capabilii levels to simulate flight. The transmission receives full torque as well as rotor
Baseline performance runs normally are conducted before and after each test. Structural load tests are performed to validate the stress model and to verify the load capacity. Thermal vacuum tests are often expensive, but they are a reliable method of determining the mechanism characteristics under environmental extremes. Life testsoften exceed the operational requirements by a factor from 1.5 to 4.0.
NOTE - Referto clause 6 for a more detaileddiscussion.
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AGMA 911-A94
This is done to ensure that all gearboxes meet the requirement. Large systems such as a robotic manipulator often can only be qualified by testing individual components due to the large size of the assembled unit. Guidance for testing can be found in Environmental Test Methods and Engineering Guidelines, MILS-1 OE. B.3.2 Test cell and installation
requirements
The condition of test cell apparatus is paramount to the successful completion of a program. Data acquisition should be complete and should be carried out in a timely manner. The following checklist may serve as a guideline for complete and orderly accrual of data: a) Pm-test: - Calibration of test apparatus and instrumentation; - Means of recording test cell temperature, humidi and atmospheric pressure; - Inspection records of all gearbox components; - Test log for proper certification of test data; - Test parameters and sequence to be cornmensurate with expected mission requirements; - List of all test equipment and facilities; - Adherence to all Safety requirements: - Review normal and emergency shutdown procedures; - Clean room conditions when required. 2) Startup: - Clean chip detectors - Heat or cool test item and oil; - Operate scavenge pumps; - Operate oil pump; - Apply load with load cell; - Check temperature, pressure, speed, vibration, flow indicators and any special test equipment. 3) During Testing: - Record all data for the established test sequence; - Proper documentation of all problems and shutdowns. 4) Post-test: -Verification of test apparatus and instrumentation calibration;
- Visual, dimensional, and non-destructive inspection record of all gearbox components; - Spectrometric analysis of the oil and chip detectors; - Evaluation of all test data in preparation for a final report.
8.4 Gearbox
Test Rigs
Accessory and actuator gearboxes may be loaded by the actual accessories or by using waterbrakes or other load absorbing devices on the mounting pads. Electric motors coupled to step up gearboxes are a common method for providing power. Variable attitude rigs are used to evaluate the effects of oil churning and scavenging during a simulated mission. In this rig the whole gearbox is mounted in a stand capable of rotation around two axes, pitch and roll, during operation. Main reduction gear driie systems would require such large loading as well as power absorption devices as to make this method of testing quite expensive. The four square system eliminates the need for a load dissipating device and reduces the size of the loading device considerably. The system is closed loop in that input and output shafts are coupled together. Figure B.6 illustrates a four square system that uses two gearboxes arranged back-to-back. One end of the two piece output (low speed) shaft is held fiied while the other end is rotated to take up backlash and then rotated until the desired torque is reached. The two halves are then locked together and connected to the prime mover through a clutch. Slip rings are used to measure torque, the clutch assures smooth acceleration of the rig tospeed, and the power required by the prime mover need only be enough to overcome the system friction which amounts to approximately 2% of the gearbox design point power. On one side of the rig, the gears will be loaded as in normal operation (test). On the other side (slave), the gears will be loaded on the coast side. Thrust load direction, for helical gears, and oil jet location may require special consideration, for the slave gearbox, when designing this type of system. The advantage of this system is the minimal amount of new hardware required, i.e. input and output shafts. The disadvantage is that the load is not variable during testing. 9393
AGMA 911-A94
inet3lkatinn ,IIY.UII~.I”II
‘J
,
Test side
T’
Slave side
c
Prime mover
Magnetic
Output (low speed) shaft Figure B.6 - Test rig arrangement, back-to-back The top view of another type of four square rig is shown as in figure 6.7. In this arrangement a single planetary reduction gearbox is coupled to a low speed gearbox on the left side and a high speed gearbox on the right. Slip ring assemblies measure both input and output torque. The gears numbered Na and Nb have the same values in both the reduction gearbox and the rig. The low speed shaft can be reduced further in speed by choosing 3Na or 4Na etc. The Input Torque Drive is located on the low speed shaft and incorporates splines that are hydraulically loaded to provide the desired torque. The advantages of this configuration are a single test gearbox is required with all gear teeth loaded and rotating as in normal operation. Also, loading may be varied during running. The disadvantage is the extra gears and attendant bearings that must be designed for each side of the rig and the increase in size of the prime mover to overcome system friction.
B.5 Production dynamic testing
gearboxes
test for larger complex gearboxes/transmissions. The complex testing may be performed in test rigs as described in B.4, or on the next higher assembly dynamic test as in the case of turbine engine accessory or speed reduction gearboxes. B.5.1 Rig acceptance tests These tests should be described in detail by acceptance test plans (ATPs). The ATP should, as a minimum, cover: a) Pretest procedures; b) Startup procedures; c) Load spectrums; d) Data to be recorded and documented; e) Inspections to be performed during test;
9 Shut down procedures; g) Post test inspections to be performed. 8.5.2 Next assembty subsystem testing
After each air vehicle gearbox/transmission undergoes final assembly, it is prudent to substantiate design performance by a dynamic test. The test
Written procedures should be used to define the test parameters as listed in B.5.1 for the gearbox/transmission being tested. This type of testing is cost effective, but does not have the risk of high
performed can be simple or complex, i.e. spin test for
disassemblycosts when infant mortality failures are
small accessory gearboxes or a mini qualification
experienced.
94
AGMA 911-A94
Test gearbox ,-
Slip ring
Slip ring
/
\-/
I I\
1
Input motor
x I
Low speed gearbox
‘7
Higk&Z$ubox Hydra&
fluid in
Figure 8.7 - Test rig arrangement, variable torque
95
AGMA 911-A94
This page is intentionally blank.
96
AGMA 911-A94
Annex C (Informative) References and Bibliography [The foreword, footnotes, and annexes are provided for informational purposes only, and should not be construed as a part of AGMA 911-A94, Design Guidelines for Aerospace Gearing.] 1. Anderson, N.E., Loewenthal, S.H.: “Comparison of Spur Gear Efficiency Prediction Methods”, Advanced Power Transmission Technology, NASA CP 2210, AVRADCOM TR 82-C-l 6, June, 1981.
9. Drago, R. J., Then Anal’icalExperimentalEvaluation of Resonant Response in High Speed, Lightweight, High/y Loaded Gearing, ASME Paper 8O-C2/DET-22, August 1980.
2. Anderson, N.E., Loewenthai, S.H., Black, J.D.: “An Analytical Method to Predict Efficiency of Aircraft Gearboxes”, ASME Journal of Mechanical Design, v. 108, Sept., 1986.
10. Lisp, T. C. and Zakrapak, J. J., ModalAnalysis of Gear Housing and Mounts, Seventh International Modal Analysis Conference, Jan. 30,1989.
3. Anderson, N.E., Loewenthal, S.H.: Efficiency of Nonstandard and High Contact Ratio Involute Spur Gears, ASME paper 84-DIET-172, presented at the Fourth ASME International Power Transmissioin and Gearing Conference, Cambridge, Mass., Oct. 1984. 4. Harris, T.A.: Rolling Bearing Analysis, Wiley, New York, 1966, pp446-450. 5. Kleckner, R.J., and Pirvics, J.: High-Speed Cy/indrical Roller Beating Ana&sis (CYBEAN) User’s Manual, SKF Report No. AL78P023, SKF Industries, Inc., (NASA Contract No. NAS3-22807), July, 1978. 6. Kleckner, R.J., Dyba, G.J., and Ragen, M.A.: Spherical Roller Bearing Anal&is (SPHERBEAN) User’s Manual, SKF Report No. AT81D007, SKF Industries, Inc., (NASA contract NAS3-22807) Feb., 1982. 7. Hadden, G.B., Kleckner, R.J., Ragen, M.A., and Sheynin, L.: System lnciuding Ball, Cylindrical, and Tapered Roller Bearings (SHABERTH) User’s Manual, SKF Report No. AT81D040, SKF Industries, Inc., (NASA Contract No. NAS#-22690) May, 1981. 8. Akin, L.S., Townsend, D.P.: Lubricant Jet Flow Phenomena in Spurand Helical Gears with Modified Addendums-for Radially Directed individual Jets, NASA TM 101460, AVSCOM TR 88-C-034, presented at the Ffih International Power Transmission and Gearing Conference, Chicago, Illinois, April, 1989.
Il. Koster, W-P., Surface /ntegrity of Machined Materials, Technical Report, AFML-TR-74-60, April 1974. 12. Fuchs, H.O., Shot Peening Stress Profi/es. 13. Lauchner, E., Westech Presentation, March 1974, Northrop Corporation, Hawthorne, California. 14. Ahmad, Aquil, Eaton Corporation. 15. Daly, J., A Concept for Using Controlled Shot Peening in Original Gear Design, American Gear Manufacturers Association Technical Paper 87FrMl3. 16. Horger, 0-J. and Lipson, C., Automotive Rear Ax/es and Means of improving Their Fatigue Resistance, American Society for Testing and Materials, Technical Publication No. 72, 1947. 17. Lowenthal, S.H.; Design of Power Transmission Shafting, NASA Report, RP-1123. 18. Townsend, D-P., and Zaretsky, E.V., Effect of Shot Peening on Su/face Fatigue Life of Canburized and HardenedAlSl9310 Spur Gears, NASA Technical Paper 2047,1982. 19. Dudley, D-W., Handbook of Practical Gear Design, McGraw-Hill, Inc., 1984. 20. Prevey, P.S., X-Ray Diffraction Residual Stress Techniques, Metals Handbook, Ninth Edition, Vol.1 0, ASM International, Ohio, 1986. 21. A. Gerve, B. Kehrwald and L. Wiesner, T. W. Conlon and G. Dearnaley, Materialscience and Engineering 69 (1966), pp. 221-225. 22. G. Hubler, I.L. Singer and C. R. Clayton, Materia/s Science and Engineering 69 (1985), p. 203.
97
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