A Comparison of the AGMA Gear Design Stresses, the Lewis Bending
Stress, and the Stresses Calculated by the Finite Element Method for
Spur Gears
by
Andrew Wright
An Engineering Project Submitted to the Graduate
Faculty of Rensselaer Polytechnic Institute
in Partial Fulfillment of the
Requirements for the degree of
Master of Engineering
Major Subject: Mechanical Engineering
Approved:
_________________________________________
Ernesto Gutierrez-Miravete, Thesis Adviser
Rensselaer Polytechnic Institute
Hartford, CT
June, 2013
(For Graduation December 2013)
i
CONTENTS
LIST OF TABLES ............................................................................................................ iv
LIST OF FIGURES ........................................................................................................... v
ACKNOWLEDGMENT .................................................................................................. vi
ABSTRACT .................................................................................................................... vii
SYMBOLS AND VARIABLES .................................................................................... viii
GLOSSARY ...................................................................................................................... x
IMPORTANT KEYWORDS ........................................................................................... xi
1. Introduction and Scope ................................................................................................ 1
2. Theory and Methodology ............................................................................................ 2
3. Analysis ....................................................................................................................... 3
3.1
3.2
Microsoft Excel Analysis ................................................................................... 3
3.1.1
Determination of the Lewis Bending Stress .......................................... 3
3.1.2
Determination of the AGMA Bending Stress ........................................ 3
3.1.3
Determination of the AGMA Contact Stress ......................................... 3
ABAQUS Finite Element Analysis .................................................................... 4
3.2.1
Lewis Bending Stress Model ................................................................. 4
3.2.2
AGMA Design Stresses Model .............................................................. 9
4. Results and Discussion .............................................................................................. 10
4.1
Comparison of Data ......................................................................................... 10
4.2
Sources of Error and Divergence ..................................................................... 12
5. Conclusion ................................................................................................................. 13
6. Appendices ................................................................................................................ 14
6.1
Microsoft Excel Analysis ................................................................................. 14
6.2
Lewis Bending ABAQUS Mesh Convergence Data ....................................... 23
6.3
AGMA ABAQUS Mesh Convergence Data.................................................... 25
ii
7. References.................................................................................................................. 26
iii
LIST OF TABLES
Table 1 – Symbols and Variables ..................................................................................... ix
Table 2 – Lewis Bending Stress by Distance through Tooth as Calculated by Abaqus .. 11
iv
LIST OF FIGURES
Figure 1 – Cut and Partitioned Pinion Tooth before Meshing ........................................... 4
Figure 2 – Boundary Conditions on the Pinion Gear Tooth .............................................. 5
Figure 3 - Coupling Constraint used to Apply Load ......................................................... 6
Figure 4 - Load Position, Magnitude, and Coordinate System .......................................... 6
Figure 5 - Visualization of Completed Pinion Tooth, with Stress Distribution ................ 7
Figure 6 – Example Graph of Stress vs Distance along Path through Gear Tooth ........... 8
Figure 7 - Path Used to Determine Bending Stresses ...................................................... 10
v
ACKNOWLEDGMENT
Type the text of your acknowledgment here.
vi
ABSTRACT
The Lewis bending stress, the AGMA bending stress, and the AGMA pitting contact
stress are calculated for a gear mesh consisting of two spur gears. Gear stresses and
factors of safety are calculated both by hand in Microsoft Excel as well as by using the
finite element method. The “Rush Gears” website has been used to generate gear CAD
files for use in the ABAQUS finite element analysis software package (see References
page). With the gear CAD files imported, ABAQUS was used to mesh, constrain, and
calculate gear stresses. The stresses calculated by these two methods will be compared
in order to determine the effectiveness of the finite element method to design gears.
vii
SYMBOLS AND VARIABLES
Symbol/Variable
Cf
CH
dg
dp
E
F
HB
I
J
KB
Km
Ko
KR
Ks
Kv
mG
ν
φ
PdG
PdP
Qv
σcG
σcP
σG
σL
σP
satG
satP
ScG
ScP
SFG
SFP
SHG
SHP
SY
Description
Units
Surface condition factor
-
Hardness ratio factor
-
Pitch diameter of the gear
Pitch diameter of the pinion
in
in
Modulus of elasticity
Face width
Brinell hardness of the gears
psi
in
HB
Pitting resistance geometry factor
Bending strength geometry factor
Rim thickness factor
Load distribution factor
Overload factor
-
Reliability factor
Size factor
Dynamic factor
-
Speed ratio
Poisson’s ratio of the gears
Pressure angle
Diametral pitch of the gear
rad
in-1
Diametral pitch of the pinion
in-1
AGMA quality factor
AGMA contact stress, gear
psi
AGMA contact stress, pinion
AGMA bending stress on the gear
Lewis bending stress
psi
psi
psi
AGMA bending stress on the pinion
Allowable bending stress number, gear
psi
psi
Allowable bending stress number, pinion
psi
Contact fatigue strength, gear
psi
Contact fatigue strength, pinion
psi
Bending fatigue failure safety factor, gear
Bending fatigue failure safety factor, pinion
psi
psi
Wear factor of safety, gear
Wear factor of safety, pinion
Yield strength of the gears
psi
viii
Equation
Used
SUT
Tg
Tp
TSG
TSP
Wt
YG
YN
YP
ZNG
ZNP
Ultimate strength of the gears
psi
Operational torque transmitted to the gear
lbf-in
Operational torque transmitted to the pinion
Stall torque of the power source, acting on the
gear
Estimated stall torque of the power source,
acting on the pinion
Tangential transmitted gear load
Lewis form factor of the gear
lbf-in
lbf-in
Stress cycle factor
-
Lewis form factor of the pinion
Pitting resistance stress-cycle factor, gear
-
Pitting resistance stress-cycle factor, pinion
-
Table 1 – Symbols and Variables
ix
lbf-in
lbf
-
GLOSSARY
x
IMPORTANT KEYWORDS
xi
1. Introduction and Scope
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provided. If your thesis is short and you want it all in one file, you may type all your
chapters in one file.
If your thesis is long, you will want to put each chapter in a separate file. Start a new
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will start with “1” in each file, but when you assemble the chapters at the end, the
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this template!
1
2. Theory and Methodology
2.1 Background of the Lewis Bending Equation and Derivation
2
3. Analysis
3.1 Microsoft Excel Analysis
3.1.1
Determination of the Lewis Bending Stress
3.1.2
Determination of the AGMA Bending Stress
3.1.3
Determination of the AGMA Contact Stress
3
3.2 ABAQUS Finite Element Analysis
This section of the report discusses in depth the two Abaqus finite element models used
to calculate the different gear stresses used in the report.
For both models, the
organization of the section will follow the different stages of the model formulation:
part(s) creation, material selection, application of boundary conditions, application of the
load, and meshing the part(s). Mesh convergence studies were carried out for both
models in order to ensure that the final mesh density resulted in sufficiently accurate
results. Refer to Appendix 6.2 and 6.3 for the raw data used to carry out the mesh
convergence for both models.
3.2.1
Lewis Bending Stress Model
To begin, the Lewis Bending Stress Abaqus model
o Describe the part
o imported RUSH gears *step file
o Removed bad geometry
o Cut into single tooth, partitioned geometry (Fig. 1)
Figure 1 – Cut and Partitioned Pinion Tooth before Meshing
o Material info
o Steel, to match excel analysis
o E = 30E6 psi, nu = 0.3
o Isotropic, elastic
4
o BC’s
o 3 total
o All DOF’s (Fig. 2)
o Explanation
Figure 2 – Boundary Conditions on the Pinion Gear Tooth
o Load
o Explain constraint (Fig. 3)
o 800 lbf load to match excel (Fig. 4)
5
Figure 3 - Coupling Constraint used to Apply Load
Figure 4 - Load Position, Magnitude, and Coordinate System
o Mesh
o Explain element type (Fig. 5)
6
o No warnings/errors
o Convergence study (Fig. 6)
o Ref. appendix with data
o Results
o Explain node path
o Screenshot of example graph
o Show
Figure 5 - Visualization of Completed Pinion Tooth, with Stress Distribution
7
Figure 6 – Example Graph of Stress vs Distance along Path through Gear Tooth
8
3.2.2
AGMA Design Stresses Model
9
4. Results and Discussion
4.1 Comparison of Data
Figure 7 - Path Used to Determine Bending Stresses
o Excel Lewis bending results
o Abaqus Lewis bending results
o Explain node path (Fig. 8)
o Speak to Table
Distance Through
Tooth (in)
0.000
35,540
0.025
11,185
0.050
5274
0.075
5811
0.094
8782
0.113
13,046
0.131
20,043
0.150
38,139
Bending Stress (psi)
10
Table 2 – Lewis Bending Stress by Distance through
Tooth as Calculated by Abaqus
11
4.2 Sources of Error and Divergence
o Mesh elements used?
o Lewis bending equation doesn’t take into account some things that Abaqus does
(stress concentrations?)
o
12
5. Conclusion
13
6. Appendices
6.1 Microsoft Excel Analysis
4.0 CALCULATION / DISCUSSION:
The following analysis calculates the AGMA design stresses for the meshing spur gears.
4.1 Input Loads
Operational Torques:
mG = dg / dp
dg and dp are the pitch diameters of the gear and the pinion, respectively.
mG [ ] =
3.0
Tg [lbf-in] =
1800
Speed Ratio (Ref. 4, Eq. 14-22)
Torque required at the gear to drive the system. This value is the same for the Excel analysis
and the ABAQUS analysis.
Tp = Tg / mG
Tp [lbf-in] =
600
Torque required at the pinion to drive the system.
4.2 Pinion, Idler, and Gear Dimensions
Pinion
F [in] =
d [in] =
Pd [in] =
dp =
PdP =
Qv [ ] =
φ [rad] =
Gear
1.25
1.5
12
1.25
4.5
12
dg =
PdG =
5
5
NG =
0.349
54
200
YG =
0.404
N[]=
HB [ ] =
NP =
0.349
18
200
Y[]=
YP =
0.309
Face width
Pitch diameter
Diametral pitch
AGMA quality
factor
Pressure angle (20 degrees)
Number of teeth
Brinell hardness of the gears (Ref. 3)
Lewis Form Factor (Ref. 4, Table 14-2)
The AGMA quality factor is also known as the transmission accuracy grade number, and is a measure of how accurate the gearing is (see Annex A, Ref. 1). Qv ranges
from 5 to 11, therefore a quality factor of 5 is a conservative estimate.
14
4.3 Material Properties and Other Input Variables
AISI 4140 steel is picked as the gear material. The following material information was retrieved from
[INSERT REF].
E [psi] =
ν[]=
Sy [psi] =
SUT [psi] =
HB [ ] =
30000000
0.3
61000
95000
197
Modulus of Elasticity of the gears (see Ref. )
Poisson's Ratio of the gears
Yield Strength of the gears (Ref. )
Ultimate Strength of the gears (Ref. )
Brinell Hardness of the gears (Ref. )
4.4 Calculation of Pitch Line Velocity
The pitch line velocity is used with the Dynamic factor, Kv, below. Increasing the pitch line velocity of the gear mesh can increase the
max stress of the gears. For this analysis, it is assumed that the gears are static to simplify the analysis. The constants affected by the
pitch line velocity become unity when the velocity is zero.
v = ωr
v [ft/min] =
0
Pitch Line Velocity of the geartrain
4.5 AGMA Bending Stress Analysis
Section 4.5 calculates the AGMA gear bending stress, the bending stress number (similar to a material strength), and a factor of safety
for bending. The calculated bending stress is based on the assumption that the gear tooth is a cantilevered plate, fixed at the base of the
tooth. This bending stress creates fatigue in the gear teeth during operation of the gear mesh. In essence, the AGMA design equations
calculate the maximum input load that the gears can withstand over the life of the gears without creating cracking. If the bending
stresses do cause cracking in the gears, they usually form at the root fillet because this is where the largest stress is. For gears with
small, thin rims, the location of the max stress can change. For this project it is assumed that the rim is of sufficient size to avoid this
situation.
Overload Factor, Ko
Ko [ ] =
1
In most practical purposes the Overlaod Factor is greater than 1 to account for momentary peak torques experienced by most
mechanically driven systems. However, in an attempt to get accurate finite element results this value was kept at 1. A constant
tangential load will be applied in the model, with no transient peaks.
15
Dynamic Factor, Kv
Kv = ((A+V0.5)/A)B
A = 50 + 56(1 - B)
B = 0.25(12 - Qv)0.66
B[]=
0.90
A[]=
55.4
Kv [ ] =
1.00
See Eq. 14-28 in Ref.
4
See Eq. 14-28 in Ref.
4
See Eq. 14-27 in Ref.
4
Size Factor, Ks
The size factor is impacted by many factors, including tooth size, diameter, face width, hardenability, and stress pattern (see Ref. 4,
section 14-10). AGMA suggests either using the equation below or simply assuming unity for this factor. I use the equation listed
because a size factor greater than 1 is conservative.
Ks = 1.192*(F*Y0.5/Pd)0.0535
See section 4.2 above for input variables.
Pinion
Ks [ ] =
KsP =
Gear
1.02
KsG =
1.03
Load Distribution Factor, Km
The load distribution factor is a ratio of the peak load to the average load applied across the entire face of the gear (Ref. 1, Annex D).
When computed analytically, this factor can be very complex. The AGMA gathered empirical data through in service gears and testing
to create the equations and variables below that are used to calculate Km.
Note: The load-distribution factor is equal to the "face load distribution factor", Cmf, under the conditions listed in section 14-11 of
Ref. 4. The gears used herein obey these assumptions, therefore Cmf is used for this factor.
Ref. 4, Eq. 14-30
16
Pinion
Gear
Cmc [ ] =
CmcP =
1
CmcG =
1
Cpf [ ] =
CpfP =
0.06
CpfG =
0.03
Cpm [ ] =
A[]=
CpmP =
Ap =
1
0.1270
CpmG =
AG =
1
0.1270
B[]=
C[]=
Cma [ ] =
Ce [ ] =
Cmf [ ] =
Km [ ] =
Bp =
Cp =
CmaP =
CeP =
CmfP =
KmP =
0.0158
-0.0001
0.15
1
1.21
1.21
BG =
CG =
CmaG =
CeG =
CmfG =
KmG =
0.0158
-0.0001
0.15
1
1.17
1.17
Eq. 14-31 from Ref. 4. Equals 1 for
uncrowned teeth
See Eq. 14-32 in Ref. 4 for 1 < F <= 17
See Eq. 14-33 in
Ref. 4
See Table 14-9 from Ref. 4. Commercial,
enclosed units
See Eq. 14-34 from Ref. 4
See Eq. 14-35 from Ref. 4
Note: Per Ref. 4, for values of F/(10d) < 0.05, F/(10d) = 0.05 is used when computing Cpf above.
Bending Strength Geometry Factor, J
The bending strength geometry factor, J, is impacted by the shape of the teeth in contact. Figure 14-6 in Ref. 4 is used to estimate this
factor based on the number of pinion and gear teeth. This factor assumes spur gears, a 20 degree pressure angle, and full-depth teeth.
Pinion
J[]=
JP =
Gear
0.32
JG =
0.40
See Fig. 14-6 in Ref. 4.
Note: Assumed that load is applied in highest point of single-tooth contact.
Stress Cycle Factor, YN
The Stress Cycle Factor alters the design stress based on the number of design stress cycles. The overall "Service Factor" used by
AGMA combines the Overload Factor, the Reliability Factor, and the Stress Cycle Factor. AGMA 2001-D04 suggests that if designers
are comfortable with the other factors in the Service Factor, unity can be used for the Stress Cycle Factor. Since this is a theoretical
problem, and the FEA software will not take into account total stress cycles, the Stress Cycle Factor has been set at 1.00.
Pinion
YN [ ] =
YNP =
Gear
1.00
YNG =
1.00
Temperature Factor, KT
KT [ ] =
1
This factor is unity unless the working temperature of the gear mesh is higher than 250 degrees
Fahrenheit (see section 14-15 of Ref. 4).
17
Reliability Factor, KR
The reliability factor takes into account normal statistical material failures that occur in material testing. Table 11 shows some common
reliability factors that were calculated from data collected by the US Navy. Unity is picked because this represents a factor in the
middle of the range.
KR [ ] =
1
See Table 11 of Ref. 1.
Rim Thickness Factor, KB
The Rim Thickness Factor is an adjustment factor that takes into account
gears with smaller "rims", the material in between the bore and the base
of the gear teeth. The factor is given in terms of the "backup ratio", the
ratio of the rim thickness of the gear to the whole depth (see Fig. X to the
right from Ref. 1). For backup ratios of greater than 1.2, the Rim
Thickness Factor becomes 1.0. For the sake of making the FEA program
simpler and getting more accurate results between the hand analysis and
the finite element, I will make the rim thickness large enough for the
backup ratio to be greater than 1.2.
KB [ ] =
Pinion
KBP =
Gear
KBG =
1.00
1.00
See Fig. B.1 from Annex B of
Ref. 1.
Gear Bending Stress, σ
Now that all the factors have been calculated, we can determine the gear bending stress. The stress is calculated below using the
equation found in Ref. 4.
Wt = 2Tp/d
Wt [lbf] =
800
σ = WtKoKvKs(Pd/F)(KmKB/J)
Wt is the transmitted tangential load going into the pinion gear.
See Ref. 4, Eq. 14-15
Pinion
σ [psi] =
σP =
Gear
σG =
29674
18
23251
4.5.1 Calculation of the AGMA Bending Fatigue Failure Safety Factor
Allowable Bending Stress Number, sat
The allowable bending stress number, sat, is
similar to a yield strength except that it goes a
step further and takes into account material
composition, cleanliness, the presence of
residual stresses, heat treatments, and materials
processing (see Ref. 1). The AGMA standard
contains tables and charts for various common
engineering gear materials with their associated
bending stress numbers. For AISI 4140, the
material of the gears, Fig. 10 is used. Grade 2,
the larger stress number, is assumed. This
grade is chosen because we will be comparing
the gears against gear models that have ideal
cleanliness and material properties.
sat = 108.6HB + 15890
Pinion
sat [psi]4 =
satP =
Gear
37284
satG =
37284
See Fig. above (from Fig. 10 of
Ref. 1)
Bending Fatigue Failure Safety Factor, SF
In engineering practice, a factor of safety is a design factor that takes into account uncertainty in the calculation of the solution. In
general, it is the ratio of the material strength of a component divided by the stress on that component. Depending on the application,
the risks involved (whether that be cost, time, or safety), and statistical randomness of the inputs, the engineer may decide to design the
component to different factors. For the purposes of the project, I chose to set a factor of safety of 1.5 as a requirement. Not only is this
a standard factor of safety for operational loading, but because AGMA has developed so many factors that make the calculated stress
more accurate we are getting results with less variance.
The AGMA standards use a factor of safety in their gear design process, which can be found in Ref. 4 (see Eq. 14-41). For this project,
the equation simplifies to sat divided by the bending stress since YN, KT, and KR are unity.
SF = [satYN/(KTKR)]/σ
Pinion
SF [ ] =
SFP =
Gear
1.3
SFG =
1.6
19
See Ref. 4, Eq. 1441
4.6 AGMA Pitting Analysis
The second important failure mode for gears is pitting, which is a surface fatigue failure that results from progressive contact stress in
the meshing gears (Ref. 4, section 14-2). Because pitting is a fatigue phenomenon, it may take many cycles to become serious enough
to result in failure of the gear system. AGMA 2001-D04 defines two types of pitting: initial and progressive. In initial pitting, small
defects are formed on the surface of the teeth in areas of high stress. These pits will, over time, correct themselves as the surrounding
high spots get smoothed out by contact with the meshing gear. For this reason, the presence of initial pitting is not a failure criteria for
gear systems. Progressiv pitting, on the other hand, does not correct itself and can occur when the stresses, lifetime cycles, or other
factors are high enough. The AGMA pitting stress equation is designed to calculate the load for which the meshing gears never
experience progressive pitting in their usage lifetime (see Ref. 1, section 4.2). This equation is based on the Hertzian contact stress
equation, modified to account for the effect of gear teeth sliding. The Elastic Coefficient, Cp, is a term that combines the elastic
material constants of the meshing gears. The equation is shown below (see Eq. 14-13, Ref. 4):
Elastic Coefficient, Cp
Cp = [1/(π((1 - νp2)/EP + (1 - νG2/EG)]1/2
Pinion
CPP =
Gear
2291
CPG =
CP [psi1/2] =
2291
See Ref. 4 Eq. 1413.
Surface Condition Factor, Cf
The surface condition factor, Cf, is a factor that takes into account surface finish effects such as cutting, lapping, grinding, or work
hardening. AGMA suggests that if the meshing gears have detrimental surface finishes caused by one of these processes the surface
condition factor should be greater than unity. For our purposes, we will make this factor unity since the finite element model will have
idealized gear surface finishes.
Cf [ ] =
1
See Ref. 4, Section 14-9
Pitting-Resistance Geometry Factor, I
As defined by AGMA, the pitting resistance geometry factor, I, evaluates how the radii of curvature of the contacting tooth profiles of
the gear mesh effects the Hertzian contact stress. Shigley's Mechanical Engineering Design, Ref. 4, provides a useful equation for
calculating I, shown below. Because the pitting resistance geometry factor only depends on the pressure angle, the load-sharing ratio,
and the speed ratio, it is the same for all gears in the mesh.
mN [ ] =
1
Load-Sharing Ratio. Equals unity for spur gears. See Ref. 4, Eq. 14-23
I = [(cosφ*sinφ)/2*mN] * (mG/(mG+1))
I[]=
0.121
See Ref. 4 Eq. 14-23
20
Gear Contact Stress, σc
The numerical value of the contact stress comes from the equation below, from Ref. 4:
σc = Cp(WtKoKvKs(Km / dP*F)(Cf / I))1/2
See Ref. 4, Eq. 14-16. Note: for the gear, dp is actually di in the equation for the
contact stress.
Pinion
σc [psi] =
σcP =
Gear
σcG =
151546
86590
Gear contact stresses
4.6.1 Calculation of the AGMA Wear Safety Factor
Similar to the value that was calculated for the bending stress, a factor of safety is calculated for the pitting stress for the pinion and
gear. Various factors must be calculated first before solving for the factor of safety.
Contact Fatigue Strength, Sc
The contact fatigue strength is calculated in a similar manner as the allowable bending stress number (or bending strength), as shown in
section 4.5.1. Since the gears are made out of AISI 4140, the contact stress number for nitrided through-hardened steel gears from
Table 3 of Ref. 1 can be used directly. Grade 2 is assumed once again in an attempt to get results that match nicely with the results of
the finite element analysis.
Pinion
Sc [psi] =
ScP =
Gear
ScG =
163000
163000
See Table 3 from Ref. 1.
Pitting Resistance Stress-Cycle Factor, ZN
Similar to how the bending stress cycle factor, YN, was handled, the pitting resistance stress cycle factor ZN will be set at unity for this
analysis. This factor alters the pitting strength that AGMA provides based on the number of lifetime stress cycles the gears will
encounter. There is no reason to change this factor from unity since there is no way to set the number of cycles in a finite element
analysis.
ZN [ ] =
Pinion
ZNP =
Gear
1.00
ZNG =
21
1.00
Hardness Ratio Factor, CH
The hardness ratio factor, CH takes into account the fact that the smaller meshing gear will see more stress cycles in the lifteime of the
gears as a result of it's smaller pitch circle. Because both gears in the mesh have the same hardness (making the ratio 1.0), however,
this factor cancels to unity.
HBP/HBG [ ] =
1.0
A' = 0 for HBP/HBG < 1.2
Hardness ratio between the pinion and the gear
Section 14-12 from Ref. 4.
CH = 1.0 + A'*(mG - 1.0)
CH [ ] =
1.0
See Ref. 4, Eq. 14-36. This factor is only used for the gear.
Wear Factor of Safety, SH
SH = [ScZNCH / (KTKR)] / σc
Pinion
SH [ ] =
SHP =
Gear
1.1
SHG =
22
1.9
See Ref. 5.2, Eq. 1442
6.2 Lewis Bending ABAQUS Mesh Convergence Data
*First column is the thru-thickness distance, and the second column is the bending stress.
Iteration 1
Iteration 2
Iteration 3
0
34858.4
0
35311.9
0
35782.1
0.007505
25571.7
0.007504
27392.3
0.007505
26555.9
0.015009
18201
0.015009
20739.2
0.015009
19130.2
0.022514
12746.3
0.022513
15352.4
0.022514
13504.4
0.030019
9207.72
0.030018
11232
0.030018
9680.22
0.037523
7585.35
0.037522
8379.7
0.037523
7656.52
0.045028
5895.34
0.045027
6354.17
0.045028
6014.95
0.052533
4877.98
0.052531
5065.21
0.052532
4996.13
0.060037
4532.59
0.060036
4510.88
0.060037
4599.05
0.067542
4859.15
0.06754
4691.16
0.067541
4823.79
0.075047
5857.29
0.075044
5605.31
0.075046
5669.93
0.082552
6827.63
0.082549
6593.34
0.082551
6714.4
0.11257
12446.5
0.090053
7840.42
0.090055
7902.6
0.120075
14327.2
0.112567
13137.9
0.112569
12330.3
0.142589
28336.8
0.142584
28786.5
0.142587
28333.3
0.150094
35796.1
0.150089
35972.1
0.150092
33884.1
Iteration 4
0
0.007505
0.015009
0.022514
0.030019
0.037523
0.045028
0.052533
0.060037
0.067542
0.075047
0.082551
0.090056
0.097561
0.11257
0.150093
Iteration 5
36895.9
26400.2
18312.4
12632.4
9585.79
7512.37
5989
5040.63
4705.5
4945.32
5759.78
6871.66
8149.43
9593.35
12899.1
37862.5
Iteration 6
0
0.007505
0.015009
0.022514
0.030019
0.037523
0.045028
0.052533
0.060037
0.075047
0.090056
0.11257
0.120075
0.127579
0.142589
0.150093
23
35548
25615.6
17968.4
12603.2
9677.79
7608.42
6014.46
5003.01
4714.17
5728.19
8163.64
12826.8
14958.9
17372.4
28288.4
37805.4
0
0.007505
0.015009
0.022514
0.030019
0.037523
0.045028
0.052533
0.060037
0.067542
0.075047
0.082552
0.11257
0.120075
0.135084
0.150094
34858.4
25571.7
18201
12746.3
9207.72
7585.35
5895.34
4877.98
4532.59
4859.15
5857.29
6827.63
12446.5
14327.2
22272.2
35796.1
Iteration 7
0
0.007505
0.015009
0.030019
0.037523
0.045028
0.052533
0.060037
0.067542
0.075047
0.090056
0.097561
0.105066
0.11257
0.120075
0.150094
Iteration 8
35446.6
25233.7
17537.5
9586.04
7576.18
6041.08
5061.26
4763.46
5014.41
5813.85
8117.91
9531.8
11177.3
13053.7
15203.6
38000.7
0
0.007505
0.015009
0.022514
0.030019
0.037523
0.045028
0.052533
0.060038
0.067542
0.075047
0.090056
0.097561
0.11257
0.120075
0.150094
24
35540.1
25166.9
17422.6
12305.7
9580.99
7571.49
6034.73
5052.7
4755.72
5008.73
5811.47
8106.95
9517.99
13046.1
15172.1
38139
6.3 AGMA ABAQUS Mesh Convergence Data
25
7. References
1.
AGMA 2001-D04, “Fundamental Rating Factors and Calculation Methods for
Involute Spur and Helical Gear Teeth.”
2.
AGMA 908-B89, “Geometry Factors for Determining the Pitting Resistance and
Bending Strength of Spur, Helical, and Herringbone Gear Teeth.”
3.
Budynas, R. G., & Nisbett, J. K. (2008). Shigley's Mechanical Engineering
Design. New York: McGraw-Hill.
4.
Cavdar, K., Karpat, F., & Babalik, F. C. (2005). Computer Aided Analysis of
Bending Strength of Involute Spur Gears with Asymmetric Profile. Journal of
Mechanical Design, 127(3), 477.
5.
Kawalec, A., Wiktor, J., & Ceglarek, D. (2006). Comparative Analysis of ToothRoot Strength Using ISO and AGMA Standards in Spur and Helical Gears With
FEM-based Verification. Journal of Mechanical Design, 128(5), 1141.
6.
Li C.-H., Chiou H.-S., Hung C., Chang Y.-Y., & Yen C.-C. (2002). Integration of
finite element analysis and optimum design on gear systems. Finite Elements in
Analysis and Design, 38(3), 179–192.
7.
Li, S. (2002). Gear Contact Model and Loaded Tooth Contact Analysis of a
Three-Dimensional, Thin-Rimmed Gear. Journal of Mechanical Design, 124(3),
511.
8.
Sfakiotakis V.G., Vaitsis J.P., & Anifantis N.K. Numerical simulation of
conjugate spur gear action. Computers and Structures, 79(12), 1153–1160.
doi:10.1016/S0045-7949(01)00014-1
9.
Zhang-Hua F., Ta-Wei C., & Chieh-Wen T. (2002). Mathematical model for
parametric tooth profile of spur gear using line of action. Mathematical and
Computer Modelling, 36(4), 603–614.
RUSH GEARS WEBSITE USED TO GENERATE CAD FILES?
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