agma 911-a94
TRANSCRIPT
-
8/16/2019 AGMA 911-A94
1/106
i
Y+ Reproduced By GLOBAL
&..= ENGtNEERlNGDOCUMENTS
--
c With The Permission Of AGMA
XB Under Royalty Agreement
AGMA911-A94
AMERICANGEARMANUFACTURERSASSOCIATION
Design Guidelines
for
Aerospace Gearing
.
AGMA INFORMATION SHEET
(This Information Sheet s NOT an AGMA Standard)
-
8/16/2019 AGMA 911-A94
2/106
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: l-55589-8294
ii
-
8/16/2019 AGMA 911-A94
3/106
AGMA 911-A94
Contents
Page
Foreword . . . . . . . . . . . . . . .
. . . . . . . . . . . . . . . . . . . . . . . . . . .
. . . . . . . . . . . . . . . . ..*...............
vi
1
1.1
1.2
2
3
3.1
3.2
4
4.1
4.2
4.3
4.4
5
5.1
5.2
5.3
5.4
5.5
5.6
5.7
6
6.1
6.2
6.3
6.4
6.5
6.6
6.7
6.8
7
7.1
7.2
7.3
7.4
7.5
8
8.1
8.2
8.3
9
9.1
9.2
Scope
.......................................................................
1
Application
...................................................................
1
References
...................................................................
1
Application
...................................................................
1
Definitions and symbols
........................................................ 2
Definitions
...................................................................
2
Symbols
.....................................................................
2
Designapproach
..............................................................
5
Design requirements and goals
... ;.
............................................
5
Identify design criieria
.........................................................
6
Preliminary design
....................................
.
.......................
8
Detail design
................................................................
12
Lubrication
..................................................................
15
Cooling vs. lubrication requirements ............................................ 15
Lubricants...................................................................l 5
Distribution systems
..........................................................
18
Lubrication system design considerations
.......................................
19
Filtration
....................................................................
21
Oiipumps
...................................................................
21
Lube system condition monitoring
..............................................
23
Environmental issues
.........................................................
24
Ambient temperature effects
...................................................
24
Ambient pressure effects
......................................................
25
Attitude effects
..............................................................
25
Contaminant effects water, corrosives, dirt, dust, and sand) ....................... 26
Vibration/Shock effects
.......................................................
26
Fire resistance requirements
..................................................
29
Electromagnetic effects
.......................................................
29
Nuclear, biological, and chemical NBC) effects
.........................
:.
.......
29
Vibration and noise
...........................................................
30
Causes of gear vibration
......................................................
30
Consequences of vibration
....................................................
31
Design
......................................................................
32
Analyzing vibration problems
..................................................
35
VibratiorYNoise reduction techniques
...........................................
37
Load Capacity
...............................................................
39
Introduction
...............................................................
..3 9
Spur, helical, and bevel gear tooth breakage and surface durability
................. 41
Spur, helical, and bevel gear scuffing scoring) -flash temperature index
............ 45
Gear materials and heat treatment
.............................................
47
Class and grade definitions
....................................................
47
Mechanical properties
........................................................
47
.
III
-
8/16/2019 AGMA 911-A94
4/106
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
10
10.1
10.2
10.3
10.4
11
11.1
11.2
11.3
11.4
12
12.1
12.2
12.3
13
13.1
13.2
13.3
Cleanliness ...............................................................
..4 8
Heat treatment ............................
...................................
48
Microstructure ...............................................................
48
Hardenability ................................................................
48
Dimensional stability
..........................................................
48
Pm-machining stock removal .................................................. 48
Ferrousgearing
..............................................................
48
Non-ferrous gearing ..........................................................
49
Material grades and heat treatment
.............................................
49
Gear surface hardening .......................................................
49
Gear through hardening
.......................................................
53
Surface treatment
............................................................
54
Introduction
.................................................................
54
Shot peening
................................................................
55
Surfacecoatings.............................................................6 0
Ion implantation of gears
......................................................
61
Manufacturing considerations .................................................. 63
Introduction .................................................................
63
Spur and helical gears
........................................................
63
Bevel gears
.................................................................
64
Stress relief treatment
........................................................
67
Gear inspect ion ..............................................................
68
General
.....................................................................
68
Spur and helical involute gears
................................................
68
Bevelgears
.................................................................
69
Rocket space gearing
....
..................................................
70
Introduction
.................................................................
70
Lubrication .................................................................. 71
Gear materia ls for space application ............................................
73
1
9
10
11
12
13
14
Symbols used in equations
.....................................................
2
Aerospace lu icant viscosities
................................................
16
Aerospace lubricant densities
..................................................
16
Aerospace lubricant pressure/viscosity coefficients
...............................
16
Aerospace lubricant specific heat values
........................................
17
Aerospacegreases
...........................................................
17
Aircraf t dry lubricants .........................................................
17
Advantages disadvantages of a common engine transmission lubrication system . 19
Particle size distribution by weight ..............................................
26
Suggested functional test levels for propeller aircraft and turbine engine equipment
... 27
Suggested functional test peak levels for equipment installed on helicopters
.........
28
Potential influence of design features on noise and vibration
....................... 35
Vibration testing
.............................................................
37
Sound and vibration reduction techniques
.......................................
39
iv
-
8/16/2019 AGMA 911-A94
5/106
AGMA Qll-AQ4
Tables,
continued
15
Typical aerospace gear materials
..............................................
49
16 Surface coatings used n aerospacegear units
.....................................
62
17
Candidate solid-film lubricants for space application
..............................
73
18 Candidate fluid lubricants for space application .................................. 74
19
Working fluid lubricants
.......................................................
74
Figures
1
2
3
8
9
10
11
12
13
14
15
16
17
18
19
20
21
22
23
24
25
26
27
28
29
Retative life as function of lambda
...............................................
7
The general parallel-axis epicyclic gear train
.....................................
9
Goodman diagram for combined load
...........................................
11
Typical aerospace lubrication system schematic
.................................
22
Spur gear pump
.............................................................
23
Vane pump
..................................................................
23
Gerotor Pump
...............................................................
23
Typical gearbox attitude limits
..................................................
25
Suggested vibration spectra for propeller aircraft and turbine eng ine equipment
...... 27
Suggested vibration spectrum for equipment installed on helicopters
................
27
Terminal-peak sawtooth shock pulse configuration and its tolerance limits
...........
28
Sound and vibration paths
.....................................................
30
Typical damping ring
.........................................................
38
Diierent methods for determining tooth root stress
...............................
40
Directions of crack propagation in gear teeth
....................................
40
Reliabilii versus number of standard deviations
.................................
46
Schematic of material ground from a gear tooth
..................................
53
Schematic of material ground from a distorted gear tooth
..........................
53
Fatigue strength in ground AISI 4349 50 HRC)
..................................
54
Example of residual stress profile created by shot peening
........................
55
Peening 1045 steel at 48 HRC with 330 shot
....................................
56
Peening 1045 steel at 62 HRC with 330 shot
....................................
56
Stress profile of carburized gear tooth root, ground and then shot peened
...........
57
Increase in fatigue resistance of spiral bevel gear
................................ 57
Fatigue tests on rear axle shafts
...............................................
57
Fatigue tests on notched shafts
................................................
57
Fatigue life comparison
.......................................................
58
Correlation of Al men intensities as indicated by arc heights of peened strips
......... 59
Heat treat coupon
............................................................
68
Annexes
A Spur gear geometry factor including internal meshes . .
. . . . . . . . . . . . . . . . . . . . . . . . . . .
75
B
Gearbox test and mission requirements
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
89
C References and bibliography
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
97
V
-
8/16/2019 AGMA 911-A94
6/106
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
-
8/16/2019 AGMA 911-A94
7/106
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 Aer ospace
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.
ASSOCIATE MEMBERS
G. Belling . . . . . . . American Pfauter
J. D. Black . . . . . . General Mo tors
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” . . . . . Kama n
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. lsabe lle . . . . . . 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
* Contributed technical material to the document.
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
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 Me mber
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 . . . . . . . . Nava l Air Propulsion Center
L. Vesey . . . . . . . . iTW/Spi roid
D. A. Wagner . . . . Genera l 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
vii
-
8/16/2019 AGMA 911-A94
8/106
This page is intentionally blank
VIII
-
8/16/2019 AGMA 911-A94
9/106
AGMA 911-A94
AGMA91 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 produc e a working gear
system, whether it be a power transmission or spe-
cial purpose mechanism. Although avariety of gear
types, such as wormgears, face gears and various
proprietary tooth forms are used in aerospaceappli-
cations, 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 guide-
lines will no longer suffice. The operat ing conditions
imposed on a high speed, high powered, transmis-
sion or actuator are quite different than those expe-
rienced by the spacecraft mechanism which must
function in a hard vacuum for long periods of time
without maintenance. This Information Sheet ad-
dresses 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 246.02A - 1983, Practice for Carburked
Aerospace Gearing.
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 pe rformed is useful in segregat-
ing the diverse gearing tasks into mechanism fami-
lies which experience similar load and environ-
mental 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 effi-
ciency;
- Helicopter transmissions. A system of
gearboxes and shafting to drive the helicopter
rotors from the engine(s);
- Mechanical interconnection between
engines allow for independent engine opera-
tion on multiingine systems;
1
-
8/16/2019 AGMA 911-A94
10/106
AGMA 911-A94
- Accessory drive gearboxes driie accessory
devices, such as generators, fuel pumps, hy-
draulic pumps, oil pressure and scavenge
pumps, blowers, alternators, etc;
- Auxiliary/secondary power units (APU/
SPU) consist of an auxiliary turbine engine inte-
grated 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 sur-
faces, 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 group ing of
power (as in rocket turbo-pump drives), and ac-
tuatortype devices which have been designed to
be compatible with the unique rigors of outer-
space environments.
These include the high
power , short life rocket applications as well as the
long life satelliie or space platform systems.
3 Definitions and symbols
3.1 Definitions.
Symbol Name
Units
C
G
Cf
CH
CL
Gl
cp
CR
G
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 fa-
miliarsymbols can be used without a careful study
of these definitions.
SI (metric) units of measure are shown in parenthe-
ses in table 1 and in the text. Where equations re-
quire 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 & pd Ks KB
*=K,K, J
. .(n)
w,Ka 1 4 KmKB
St=--
K~K~ mF
. .(llM)
J
The second expression uses SI units.
Table 1 - Symbols used in equations
Center distance
Application factor for pitting resistance
Surface condiiion factor for pitting resistance
Hardness ratio factor for pitting resistance
Life factor for pitting resistance
Load distribution factor for pitting resistance
Elastic coefficient
Reliability factor for pitting resistance
Size factor for pitting resistance
in (mm)
----
a
w
First
Reference
equation paragraph
9
8.2.2
12
8.2.2
12
8.2.2
18
8.2.8
18
8.2.8
12
8.2.2
12
8.2.2
18
8.2.8
12
8.2.2
2
-
8/16/2019 AGMA 911-A94
11/106
AGMA 91 A94
Table 1 (coMwo”
Symbol
Name
Units
First Reference
equation 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,
Reduced modulus of elasticity
IWir?(N/mm2) 4 4.2.4
F
Face width
in (mm)
9
8.1.2
Fe
Effective or net face width
in (mm)
12 8.2.2
H
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 )
(mm)
11 M 8.2.1
m
Gear ratio (never less than 1 O)
----
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
-
8/16/2019 AGMA 911-A94
12/106
AGMA 91 -A94
Table 1 (concluded)
Symbol
pd
sac
sat
SC
St
vc
swt
Till
T
ut
82
tfl
tM
u
u’
Ve
W
W’
WVr
wt
Xl-
a
h
CL
PO
V
Pn
(Ja
Ul,(x
Name
Diametral pitch ( = 25.4/m )
Reliabilii constant
Allowable contact stress number
Allowable bending stress number
Contact stress number
Bending stress number
Working contact stress number
Working bending stress number
Inlet oil temperature
Outlet oil temperature
Contact temperature
Flash temperature
Bulk temperature
Speed parameter
Average rolling speed
Entraining velocity
Load parameter
Unit tangential load
Normal unit load
Tangential tooth load
Load sharing factor
Pressure viscosity coefficient
Specific film thickness
Viscosity
Absolute viscosity
Coefficient of variation or standard deviation
Normal relative radius of curvature
Composite surface roughness
Surface roughness of pinion, gear
Unite
in-l
----
lb/in* (MPa)
lb/in* ( MPa)
lb/in* ( MPa)
lb/in* (MPa)
lb/in* (MPa)
lb/in* ( MPa)
OF “C)
OF “C)
“F (“C)
OF “C)
OF “C)
----
inlsec (mm&c)
in/s (m/s)
----
lb/in (N/mm)
lb/in (N/mm)
lb 0’4
----
in*/lb (l/MPa)
inlin (mm/mm)
reyns (kPa s)
reyns (kPa s)
----
in (mm)
tin @ml
crin tw)
First Reference
iquatior
paragraph
11 8.2.1
I4 8.2.7
18 8.2.8
13 8.2.7
I2
8.2.2
II 8.2.1
I8 8.2.8
I3 8.2.7
8 5.1.2
8
5.1.2
I9 8.3.1
I9 8.3.1
I9
8.3.1
6 4.2.4
1 4.2.4
6
4.2.4
7
4.2.4
I 4.2.4
7
4.2.4
I1 8.2.1
7
4.2.4
5
4.2.4
2
4.2.4
1
4.2.4
6
4.2.4
I4
8.2.7
I
4.2.4
2 4.2.4
3
42.4
4
-
8/16/2019 AGMA 911-A94
13/106
AGMA 911-A94
4 Design approach
4.1J Maintainability
4.1 Design requirements and goals
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 maxi-
mum 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 evels must be specified
for a new gearbox system.
Guidelines for field service work, space require-
ments, and tool limitations must be specified early in
the project.
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 establi shed 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 parame-
ters. 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 l anding 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 con-
siderations.
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
-
8/16/2019 AGMA 911-A94
14/106
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, proper-
ties of materials used, and the shape of the tooth
surfaces near the point of contact before the load
transfer begins.
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.
4.2.1.2 Actuator gearing
Actuator gears are subject to “holding” loads which
are static loads. These loads occur in systems such
as aircraft flap drive systems, winches, and space-
craft 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 accommo-
date 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 de-
formation 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 condi-
tions 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 hard-
ened 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
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.
Allowable bending atigue limits are given in ANSI/
AGMA 2001-B88.
6
-
8/16/2019 AGMA 911-A94
15/106
AGMA 911-A94
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 in-
dependent of gear design and operating conditions.
Appendix A of ANSVAGMA 2001-888 defines
scuffing as related to the instantaneous tempera-
ture rise on tooth surfaces caused by frictional heat.
The equations which define the surface tempera-
ture 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 condi tions, 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
ANSVAGMA 2001-B88.
Maximum values
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.
4.2.4 Lubricant film thickness
Lubricant film thickness has received ever-increas-
ing attention since the time it was introduced by
Martin in London Engineering in 1914.
H= .
896 ’ “h
H = ,,,,g:”
” pn
W’
where
H is oil film thickness, in (mm);
p is viscosity, reyns (kPa s);
U’ is average rolling speed, inls (mm/s);
h is normal relative radius of curvature, in
(mm):
d 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, Dow-
son, 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 sur faces, 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 opera-
tion of the gears.
NOTE- issupplanting
ms as a way of describ-
ing roughness. Both terms are still in use but are
not equivalent.
2.2
-i
Aerospace gears
//
g
a,
5
, A Bkaririgs
/
4 -6 1
2 4
ecific film thickness, h
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@
-
8/16/2019 AGMA 911-A94
16/106
AGMA 911-A94
Further studies by NASA simplified the general
accessory gears. If equation 4 is used for power
equation to the form presented in Appendix A of
gearing without the previously noted enhance-
ANSVAGMA 2001-B88 for dimensionless mini-
ments, the definition of when boundary lubrication
mum film thickness:
occurs may be as low as R = 0.2 to E.= 0.4.
H
GO54 uo.70
mh = 2.65
.*
WO.13
p-(4)
where the following are dimensionless parameters:
materials parameter, G;
G= a,?
. .(5)
speed parameter, V;
PO Ve
u= -
2w+a
load parameter, W,
e(7)
where
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);
Xr load sharing factor;
wr normal unit load, lb/in (N/mm).
The following enhancements may be added to the
calculation as follows:
- Transient squeeze film effects from change in
entrainment velocity, surface geometry and dy-
namic 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 co-
efficients when oil jet cooling is used for minimi-
zation 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.
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
The areas of concern during the preliminary phase
of aerospace gearbox design consist primarily of
performance, cost, configuration and packaging.
45.1 Configuration study
In the preliminary design stage, it is generally
necessary to lay out various gea rbox configurations
which meet the basic speed, power, and ratio
requirements. These configurations can be com-
pared 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.
4.3.1.1 Gearing
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 compli-
cated by pitch change mechanisms, accessories,
overrunning clutches, engine air intake, etc.
Possible configurations include:
- 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
The relative film thickness, as calculated using
equation 4 for ZYZ~
has been derived and used
successfully using narrow face width gears such as
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
8
-
8/16/2019 AGMA 911-A94
17/106
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 genera l 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 in-
ternal sun against rotation its pitch circle has zero
angular velocity and the remaining three compo-
nents, 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 bal-
anced 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
-
8/16/2019 AGMA 911-A94
18/106
AGMA 911-A94
free from high centrifugal bear ing 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 mem-
ber 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 arrange-
ments. 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
cautioned however, that some compound planetary
variations exhibit very poor transmission efficiency
due to high effective pitch line velocity in the h igh
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 redun-
dant 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 dedi-
cated 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 acces-
sories and the position of the gearbox input drive.
When positioning the accessory gearbox, consid-
eration needs to be given to the overall configura-
tion 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 for specific
arrangements.
45.2 Gear sizes
There usually are two modes of operation which
size gears as follows:
- Start up conditions;
10
-
8/16/2019 AGMA 911-A94
19/106
-
8/16/2019 AGMA 911-A94
20/106
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.
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 rad ial 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 mem-
bers could bind leading to increased operating
stresses.
A flexible spline is not held radially by a diametral fit.
This pe rmits 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
which can accelerate wear and thus will require
periodic maintenance.
Flexible splines used as
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.
4.3.4.2 Bearings
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.
As the bearing size increases it is generally more
difficult to obtain calculated life due to changes in
preload caused by mounting thermal and centrifu-
gal 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 magneti-
cally loaded face seals bore rubbing seals laby-
rinth 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
Detail design of a geared system requires accurate
evaluation of dynamic gear tooth loads caused by
12
-
8/16/2019 AGMA 911-A94
21/106
AGMA 9ll-A94
load transfer from one mesh to another and
4.4.1 Finite element modeling considerations
momentary overloads caused by system reso-
nance. 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 signifi-
cantly 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 genera l use.
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 dis-
cussed earlier.
4.42 Tooth bending and contact stress con-
siderations
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 calcula-
tions and limits are discussed in clause 8.
Gear stresses are a valuable design tool in deter-
mining 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 endur ance
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.
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.).
- Pitting isvery much affected by lubrication con-
ditions;
- There is no pitting endurance limit. S-N
diagram does not become asymptotic.) The
allowable stress used for design purposes con-
siders such items as the number of cycles and
the types of material and oil used.
Load distribution is influenced by the above factors
and is non-uniform along the contact lines of
meshing gears. To determine tooth load distribu-
tion, tooth and r im deflections are required. These
deflections vary with the load position and affect the
dynamicsand tooth root stress as the tooth rotates
through the entire mesh.
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
metal-to-metal contact. When this favorable
situation was obtained, the gear or the bearing
13
-
8/16/2019 AGMA 911-A94
22/106
AGMA Qll-A94
could either carry more load without pitting or run for
a longer time without pitting at a given load.
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) ycles 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.
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 a reas 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 Reg ime Ill. Helicopter gears are in the
high speed gear region at the input sections of the
gearbox.
Definition of endurance limits and regime of lubrica-
tion 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
finish, and designed to controlled surface finish and
waviness.
Aircraft bearings are typically AFBMA grade 5 to 7
or better, selectively designed to meet performance
requirements.
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 environ-
mental conditions outlined in clause 6.
Vibrations, fire resistance, weapons effects, emis-
sions, and attitude are also integral parts of the
above defined tests.
* Numbers in brackets ] refer to references listed in Annex C.
14
-
8/16/2019 AGMA 911-A94
23/106
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 lu-
brication include: scuffing, micropitting and spalling.
5.1 .l Elastohydrodynamic (EHD) lubrication
and lambda ratio
The thickness of the protective EHD oil film can be
calculated using the techniques described in ap-
pendix 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 a t 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 lubrica-
tion 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 entire ly 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 Lubr icants
52.1 Liquid lubricants
Liquid lubrication predominates in the aerospace in-
dustry 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 pre-
vent mixing of two diierent types of oil. These lubri-
cants were formulated to meet criteria such as cold
flow/cold start requi rements, high temperature li-
mitations, 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 butad-
diiives beneficial to the transmission are included.
Tables 2 through 5 list pertinent proper ties of the
most commonly used aircraft lubricants today.
15
-
8/16/2019 AGMA 911-A94
24/106
AGMA 91%A94
Table 2 - Aerospace lubricant viscosities
Tern1
OF
ViSCOS
MIL-L-7808’
I,csf
DOD-L-85734*
Dexron II3
-
-
-
2.23
-
2.8
-
7
5.00 to 5.50
-
-
42
25.00 min
-
c9500
20000
-
-
lrature
“c
MI L-L-23699’
204 1.25
177
1.63
160 2.00
100 5.00
98.9
5.ooto5.50*
40
25.00
37.8 25.00 min*
-40 13000max*
-
400
350
320
212
210
104
100
-40
-65
1.00
1.25
1.47
3.00min”
-
12.00
-
2000
13OOOmax
dotes-
Reference - 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
OF
, s/ml
DOD-L-85734*
Densi
MlL-L-7808’
392 200
320 160
302 150
212 100
104 40
60
16
Dexron II3
IL-L-23699’
0.86
0.81
0.89 0.84
0.90
0.85
0.94
0.89
0.98 0.93
-
-
-
0.87
Jotes -
Reference - AFAPL-TR-71-35
1Exxon Datasheet - ET0 25
I General Motors Dexron II Specification
Table 4 - Aerospace lubricant pressure-viscosity coefficients
Temperature
“F “C
400 204
350 177
320 160
212 100
104 40
Pressure-viscosity coefficient,
(in2/lb)x10000[(mm2/N)x10 000]
MIL-L-23699*
MIL-L-7808G*
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)
l Reference AFAPL-TR-75-26
16
-
8/16/2019 AGMA 911-A94
25/106
AGMA 911-A94
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 com-
mon. The most common greasesand their uses are
listed in table 6.
5.2.3 Dry lubricants
Dry lubricants are widely used in spacecraft sys-
tems 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.
Table 5 - Aerospace lubricant
specific heat values
Specific heat,
btu’lb OF kJ/(kg OK)]
MIL-L-23699 or
MIL-L-7808G*
Temperature
“F “C
400 204
350 177
t
300 149
250 121
200 93
150 66
100 38
0.562 (2.35)
0.551 (2.31)
0.538 (2.25)
0.524 (2.19)
0.508 (2.13)
0.486 (2.03)
0.464 (1.94)
’ Reference AFAPLR-T 75-26
Table 6 - Aerospace greases*
ML Specification
Description
Application
MIL-G-6032
Oil resistant grease
Tapered plug valves, gaskets
MIL-G-21164
Molybdenum disulfide
Splines, sliding steel surfaces
MIL-G-23827
Gear and actuator grease Bearings, gears, etc.
MIL-G-25013
Aircraft bearing grease
Ball and roller bearings to 200 000 DN
MlL-C-38220
High speed bearing grease Ball and roller bearings to 400 000 DN
MIL-L-27617
Oil resistant grease
Tapered plug valves and gaskets
MlL-G-46006
Aircraft grease
Driveshaft couplings
MIL-G-81322
General purpose grease
Bearings and gearboxes
MIL-L-81827
High load capacity grease
Splines and bearings
MIL-G-83261
Extreme pressure grease
Gearboxes, actuators
MIL-G-83363
Helicopter transmission grease
Tail rotor and intermediate gearboxes
l 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).
Table 7 - Aircraft dry lubr icants*
MIL Specification Description Application
m-G-659
Graphite
Dry lubricant or mix with oil
MIL-M-7866
Molybdenum disulfide
Threads, gears
MIL-L-8937
Corrosion inhibiting
Gears, flap hinges
MIL-L-23398
Air drying solid film
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
-
8/16/2019 AGMA 911-A94
26/106
-
8/16/2019 AGMA 911-A94
27/106
AGMA 911-A94
stated above for bear ings and spiines they are also
less vulnerable in combat situations.
Flap actuator gear drives are another common
aerospace application for grease lubrication.
5.3.5 Powder lubrication
Powder lubrication is being investigated for very
high temperature applications where only dry lubri-
cants can survive. in these systems dry lubricants
are blown into bear ings and gears as an aerosol to
lubricate and cool the system.
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 ad-
vantages and disadvantages of each type of sys-
tem.
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 opera-
tion 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
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.
5.4.3 Oil deaeration
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.
5.4.4 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 require-
ment frequently results in multiple oil scavenge
pumps with oil pick-up passages placed n 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
Increased weight - redundant parts
Engine oil must be used
Transmission oil can be used
Transmission failure can affect engine
Transmission failure is self contained
External plumbing required
Self containedplumbing
19
-
8/16/2019 AGMA 911-A94
28/106
AGMA 911-A94
Scavenge oil pumps are typically rated at two to
When taking an oil sample it is good practice to
three and sometimes more) times the supply pump
clean the tap priorto taking the sample and to let the
flow rate. This ensures that oil will not build up in
oil flow for several minutes before taking the actual
critical areas under adverse operating conditions.
sample.
5.4.7 Fill/drain considerations
ne 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 p lacement 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