Wear of Spur Gears Having a Dithering Motion and

Lubricated With a Perfluorinated Polyether Grease

NASA/TM—2007-215008

December 2007

Timothy Krantz

U.S. Army Research Laboratory, Glenn Research Center, Cleveland, Ohio

Fred Oswald

Glenn Research Center, Cleveland, Ohio

Robert Handschuh

U.S. Army Research Laboratory, Glenn Research Center, Cleveland, Ohio

ARL–TR–4124

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Wear of Spur Gears Having a Dithering Motion and

Lubricated With a Perfluorinated Polyether Grease

NASA/TM—2007-215008

December 2007

Timothy Krantz

U.S. Army Research Laboratory, Glenn Research Center, Cleveland, Ohio

Fred Oswald

Glenn Research Center, Cleveland, Ohio

Robert Handschuh

U.S. Army Research Laboratory, Glenn Research Center, Cleveland, Ohio

ARL–TR–4124
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Level of Review: This material has been technically reviewed by technical management.
NASA/TM—2007-215008 1

Wear of Spur Gears Having a Dithering Motion and Lubricated 

With a Perfluorinated Polyether Grease 

Timothy Krantz 

U.S. Army Research Laboratory 

Glenn Research Center 

Cleveland, Ohio 44135 

Fred Oswald 

National Aeronautics and Space Administration 

Glenn Research Center 

Cleveland, Ohio 44135 

Robert Handschuh 

U.S. Army Research Laboratory 

Glenn Research Center 

Cleveland, Ohio 44135 

Abstract 

Gear contact surface wear is one of the important failure 

modes for gear systems. Dedicated experiments are required 

to enable precise evaluations of gear wear for a particular 

application. The application of interest for this study required 

evaluation of wear of gears lubricated with a grade 2 

perfluorinated polyether grease and having a dithering 

(rotation reversal) motion. Experiments were conducted using 

spur gears made from AISI 9310 steel. Wear was measured 

using a profilometer at test intervals encompassing 10,000 to 

80,000 cycles of dithering motion. The test load level was 

1.1 GPa maximum Hertz contact stress at the pitch-line. The 

trend of total wear as a function of test cycles was linear, and 

the wear depth rate was approximately 1.2 nm maximum 

wear depth per gear dithering cycle. The observed wear rate 

was about 600 times greater than the wear rate for the same 

gears operated at high speed and lubricated with oil. 

 

Introduction 

Gear contact surface wear is one of the important failure 

modes in gear systems. Wear and the associated material loss 

can lead to structural failure (gear tooth fracture). Wear can 

also lead to changes in vibration and noise behavior (ref. 1 

to 3). In addition, wear can change the patterns of gear 

contact such that the altered load distributions and contact 

stresses will accelerate the occurrence of other failure modes 

such as pitting and scoring (ref. 4). Gear wear debris can also 

be detrimental to the performance of bearings or other 

components of a drive system (ref. 5). 

The study of wear is becoming one of the emerging areas of 

gear research. A number of recent wear modeling efforts (refs. 6 

to 10) form a solid foundation for studying gear wear. The 

common thread to these studies is that all use the well-known 

Archard’s wear model (ref. 11) in conjunction with a gear contact 

model and relative sliding calculations. Archard’s wear equation 

can be expressed for a local point on one of the contacting gear 

surfaces as dh/ds = kP where k is an experimentally determined 

wear coefficient, h is the wear depth accumulated, P is the contact 

pressure, and s is the sliding distance between the mating surfaces 

at the point of interest. From this equation, to calculate the wear 

depth h, the contact pressure P and the sliding distance s must be 

determined. Flodin and Andersson (refs. 6 to 8) and Bajpai et al. 

(ref. 9) proposed wear models for spur and helical gears, and the 

focus was to determine P and s. The tooth contact pressures P

were computed in these models using either simplified Hertzian 

contact (refs. 6 to 8) or boundary element (ref. 9) formulations 

under quasi-static conditions. Sliding distance s calculations were 

determined from gearing kinematics, and Archard’s wear model 

(ref. 11) was used with an empirical wear coefficient to compute 

the surface wear distribution. 

Perhaps the most significant shortcoming of the gear wear 

prediction models discussed above is that they consider wear 

depth to be a function of only two parameters, contact pressure, 

P and sliding distance, s. All other influences, such as surface 

material, surface roughness and lubrication at the contact 

interface, are all accounted for by the wear coefficient k

(ref. 12). When all of the parameters including lubricant type, 

temperature, flow rate, gear material composition, surface 

topography, and surface hardness are consistent, then it is 

possible to define K using a small number of controlled 

experiments (ref. 9). However, if all of the stated properties are 

not consistent, then determining a wear rate coefficient is a 

challenging task. In addition, the influence of such parameters 

on wear cannot be described by these models. As such, 

dedicated experiments are often required to enable accurate 

predictions of gear wear for a particular application. 

Perfluorinated polyether greases are commonly used for 

space mechanisms. Discussions in support of mechanism 

maintenance (ref. 13) have highlighted the need for data to 

better understand the wear and behavior of mechanical 

components operated using such grease. Some qualitative 

results of gear wear experiments were previously reported 

(ref. 14). To quantify wear rates for gears lubricated with 
NASA/TM—2007-215008 2

perfluorinated polyether grade 2 grease, spur gear experiments were completed. 

Test Methodology 

Test Rig 

The experiments were conducted using the NASA Glenn 

Research Center Spur Gear Fatigue Test Rigs. These test rigs 

have been used for more than 30 years to test oil lubricated 

spur gears, with emphasis on studying contact fatigue 

(spalling, pitting, and micropitting). The test rig, as shown in 

figure 1(a), uses the four-square (torque-regenerative) principle of applying test loads, and thus the motor needs to 

overcome only the frictional losses in the system. The test rig 

is belt driven using a variable speed electric motor. A 

schematic of the loading apparatus is shown in figure 1(b). 

Hydraulic oil pressure and leakage replacement flow is 

supplied to the load vanes through a shaft seal. As the oil 

pressure is increased on the load vanes located inside one of 

the slave gears, torque is applied to its shaft. This torque is 

transmitted through the test gears and back to the slave gears. 

In this way power is circulated, and the desired load and 

corresponding stress level on the test gear teeth may be 

obtained by adjusting the hydraulic pressure. 

Figure 1 depicts the spur gear rig as has been used for tests 

operated at 10,000 rpm for the purpose of evaluating the 

fatigue lives of oil lubricated gears. The test setup as used for 

the grease tests reported herein differed from the depiction of 

figure 1 in two important ways. Figure 1 illustrates the test 

gears operating with faces offset. The face-offset condition is 

used to concentrate the Hertz contact stress as is desired for 

accelerated life testing of high cycle fatigue. For the grease 

lubricated gear testing, the gears were operated with zero 

offset (full faces in contact with each other). Also, figure 1 

depicts pressurized labyrinth seals on the two shafts. For the 

grease testing reported herein, lip seals were used on the two 

shafts to prevent leakage of the slave gear lubricating oil to 

the grease lubricated test gear section. The lip seals have been 

used with much success on these rigs to maintain zeroleakage even for speeds of 10,000 rpm. 

For some applications, gear teeth will operate both as a 

driving and as a driven member depending on the motions 

and torques applied to the machine at any given instant. 

Hall et al. (ref. 13) provide an example of an actuator that 

operated in a dithering mode whereby a gear rotated with a 

back and forth motion with a nominally constant torque 

(fig. 2). For such operation, the gears can create wear debris 

on an approach path and then entrain the debris into the gear 

contact on the return path. For the subject experiments, the 

gear test rig was operated in such a dithering mode. To create 

the dithering motion, a four-bar linkage mechanism was 

attached to the drive shaft of the test rig (fig. 3). The four-bar 

linkage was designed to allow for adjustments to the range of 

motion. For the tests reported herein a full rotation of the 

electric motor rotated the gears through ±28° of rotation. This 

range of motion ensured that one tooth pair would operate for a 

full gear tooth engagement cycle for each forward and return 

stroke of the mechanism. The frequency of motion was 4 Hz 

(4 dithering cycles per second). 

Figure 1.—NASA Glenn Research Center gear fatigue test 

apparatus. (a) Cutaway view. (b) Schematic view. 

(a) 

(b)

Figure 2.—Illustration of the reversal of motion and constant 

direction of torque used for wear testing. (a) Forward stroke 

of dithering motion. (b) Reverse stroke of dithering motion. 
NASA/TM—2007-215008 3

4-bar mechanism test gears angular motion sensor

Figure 3.—Spur gear test rig configured for grease lubricated testing using a dithering motion. 

Test Gears and Lubricant 

The test gears used for this work were manufactured from 

a single lot of consumable-electrode vacuum-melted (CVM) 

AISI 9310 steel. The nominal chemical composition of the 

AISI 9310 material is given in table 1. The gears were case 

carburized and heat treated according to table 2. The nominal 

properties of the carburized gears were a case hardness of 

Rockwell C60, a case depth of 0.97 mm (0.038 in.), and a core 

hardness of Rockwell C38. The dimensions for the test gears 

are given in table 3. The gear pitch diameter was 89 mm 

(3.5 in.), and the tooth form was a 20° involute profile 

modified to provide a tip relief of 0.013 mm (0.0005 in.) 

starting at the highest point of single tooth contact. The gears 

have zero lead crowning but do have a nominal 0.13-mm- 

(0.005-in.-) radius edge break at the tips and sides of the teeth. 

The gear tooth surface finish after final grinding was specified 

as a maximum of 0.406 μm (16 μin.) rms. Tolerances for the 

gear geometries were specified to meet AGMA (American 

Gear Manufacturers Association) quality level class 11. Gears 

were lubricated using a single container of grade 2 

perfluorinated polyether grease. The grease is qualified for use 

on space mechanisms. 

TABLE 1.—NOMINAL CHEMICAL COMPOSITION 

OF AISI 9310 GEAR MATERIAL 

Element wt% 

Carbon 0.10 

Nickel 3.22 

Chromium 1.21 

Molybdenum 0.12 

Copper 0.13 

Manganese 0.63 

Silicon 0.27 

Sulfur 0.005 

Phosphorous 0.005 

Iron balance 
NASA/TM—2007-215008 4

TABLE 2.—HEAT TREATMENT FOR AISI 9310 GEARS 

Step Process Temperature Time, hr 

 K oF 

1 Preheat in air — — — 

2 Carburize 1,172 1,650 8 

3 Air cool to room 

temperature 

— — — 

4 Copper plate all over — — — 

5 Reheat 922 1,200 2.5 

6 Air cool to room 

temperature 

— — — 

7 Austentize 1,117 1,550 2.5 

8 Oil quench — — — 

9 Subzero cool 180 -120 3.5 

10 Double temper 450 350 2 each 

11 Finish grind — — — 

12 Stress relieve 450 350 2 

 

TABLE 3.—SPUR GEAR DATA 

[Gear tolerance per AGMS class 11] 

Number of teeth 28 

Module, mm 3.175 

Diametral pitch 8 

Circular pitch, mm (in.) 9.975 (0.3927) 

Whole depth, mm (in.) 7.62 (0.300) 

Addendum, mm (in.) 3.18 (.125) 

Chordal tooth thickness 

reference, mm (in.) 

4.85 (0.191) 

Tooth width, mm (in.) 6.35 (0.25) 

Pressure angle, deg. 20 

Pitch diameter, mm (in.) 88.90 (3.500) 

Outside diameter, mm (in.) 95.25 (3.750) 

Root fillet, mm (in.) 1.02 to 1.52

(0.04 to 0.06) 

Measurement over pins, mm 

(in.) 

96.03 to 96.30

(3.7807 to 3.7915) 

Pin diameter, mm (in.) 5.49 (0.216) 

Backlash reference, mm (in.) 0.254 (0.010) 

Tip relief, mm (in.) 0.010 to 0.015

 (0.0004 to 0.0006) 

Test Procedure 

The test gears were cleaned to remove a rust-preventative 

preservative, assembled on the test rig, and grease was 

applied. The gears were tested with the tooth faces fully 

engaged (the faces were not offset as depicted in Fig. 1). Tests 

were run at a frequency of 4 full dither cycles per second. All 

tests were conducted with a hydraulic pressure of 1.72 MPa 

(250 psi) applied to the loading device. The torque produced 

for such a hydraulic pressure was verified both before and 

after testing to be 68 N-m (51 ft-lb). The applied torque 

resulted in a contact condition of 1.1 GPa (160 ksi) maximum 

Hertz contact stress at the pitch-line. The Hertz stress just 

stated is an idealized stress index assuming static equilibrium, 

gears meshing at the pitch-point position, perfectly smooth 

surfaces, and an even load distribution across the measured 

active face width. 

The active face width was measured using a profilometer 

tracing in the lead direction. The profile trace was used to 

determine the effect of the edge break radius on the active face 

width. Three pairs of gear teeth were tested. For each tooth 

pair, the testing was stopped three times to make tooth profile 

measurements. The test matrix of running times (dithering 

cycles) is listed in table 4. Prior to measuring the tooth 

profiles, grease and wear debris were removed from the teeth 

using soft-tipped tools. Chemicals were not used to assist the 

removal of the grease and wear debris to avoid contamination 

of the tooth surface that was to undergo additional testing. 

 

TABLE 4. —TEST MATRIX FOR SPUR GEAR 

WEAR TESTING SHOWING CYCLE COUNT 

WHERE TESTING WAS STOPPED FOR 

WEAR MEASUREMENTS 

Tooth Pair Accumulated Number 

of Dithering Cycles 

B 5,000 

B 10,000 

C 10,000 

A 20,000 

C 20,000 

A 40,000 

B 40,000 

A 80,000 

C 80,000 

Measurement of Wear 

Wear was determined by measuring the gear tooth profiles 

prior to testing, then again measuring the tooth profiles after 

testing, and comparing the data. The teeth were measured 

using a stylus profilometer with a small-bore diamond-tipped 

stylus having a 2 μm 90° tip. Three traces were obtained for 

each gear for each test condition, the three traces at nominally 

equal spacing across the face width. The profilometery data 

were not filtered (other than the mechanical filtering inherent 

to the stylus-type measurement technique). The profiles 

measured were of 7.2 mm length, tracing from below the true 

involute form radius in the root area of the gear to beyond the 

tip and across the topland of the tooth. The concept of the 

wear measurement procedure involved overlaying traces of the 

new gear and worn gear, with any differences attributed to 

wear. To assist with the overlaying procedure, the gears were 

mounted on the measuring machine using a fixture employing 
NASA/TM—2007-215008 5

a gear rack (fig. 4). The fixture was used to fix the orientation 

of the gear relative to the measuring machine coordinate axes. 

With such a fixture, profiles could be overlaid after translation 

of one trace relative to the other. Rotations of coordinate 

systems were not necessary. A typical set of raw data plots for 

a gear tooth, before and after testing, is provided in 

figure 5(a). 

To determine wear from profilometer traces, it was 

necessary to first overlay the two traces and then subtract one 

from the other. The overlaying of traces was done using a 

semi-automated procedure. The first step was to visually 

translate the data for profile of the “worn” tooth until the plots 

of the two profiles were closely matched in the root and 

topland parts of the profiles (that is, those locations with no 

contact and no wear). After this initial manual operation of 

overlaying was completed, a computer algorithm was then 

used to optimize the translation of the trace. The optimization 

was done to minimize the sum of the squares of the distances 

between the nearest-neighbor datapoints on the two traces in 

the regions of the root and topland that were not in contact. A 

typical dataplot for two traces after optimizing the translation 

of the worn trace is provided in figure 5(b). Subtracting the 

“new” profile data from the “worn” profile data point by point 

yielded the wear of the tooth as a function of position along 

the tooth (fig. 5(c)). 

Wear amounts were characterized and summarized as 

follows. For each wear profile the maximum depth of wear 

and the area under the wear curve (by numerical integration) 

were determined. The wear area can be considered as a 

measure of the volume of material removed per unit facewidth 

of the gear tooth. The wear depth and wear area data were fit 

to regression models. To complete the regression and 

accompanying statistical analysis, the data were first 

transformed as X = (number of dithering cycles)1/2 and Y = 

(wear)1/2. Such a transformation was found to provide a 

regression model having X proportional to Y and having 

residuals with normal distributions and equal variance for all 

tested values of X. Two-sided point-by-point 95 percent 

tolerance intervals at 95 percent statistical confidence were 

determined using the method of Hahn and Meeker (ref. 15). 

Figure 4.—Gear mounted on a profilometer using a 

gear rack to locate and orient the tooth. 

01234567

profile height (mm)

-0.8

-0.6

-0.4

-0.2

0.0

0.2

0.4

before testing

after testing

01234567

profile height (mm)

-0.8

-0.6

-0.4

-0.2

0.0

0.2

0.4

before testing

after testing

position along gear tooth (mm)

01234567

wear depth (mm)

0.02

0.04

0.06

0.00

region having material loss

(a)

(b)

(c)

root tip 

Figure 5.—Method for determining wear from profilometer data. 

(a) Raw traces of the profilometer data. (b) Profilometer data 

after X and Y translations of the second trace to matchup 

profiles in the tooth root and topland regions. (c) Wear depths 

as obtained by the difference of the two traces of plot (b). 

Test Results 

To quantify wear rates for gears lubricated with 

perfluorinated polyether grade 2 grease a set of spur gear 

experiments were completed. The wear amounts were 

characterized by (a) the maximum depth of wear and (b) the 

total volume of material removed per unit face width (termed 

wear area). Wear profiles for all test conditions are provided in 

reference 13. The gear teeth developed unique wear patterns. 

That is, there were significant differences in wear profile 

amount and shapes from one tooth to another. It is speculated 

that the wear patterns are determined by the detailed geometry 

and conditions of the mating teeth including all of form, 

waviness, roughness features, and hardness. 

A photograph showing the typical condition of a worn 

tooth is provided in figure 6. The visual appearance of the 

tested surface differed from tip to root, likely depending on the 

amounts of relative sliding. In areas of high relative sliding 

(toward the tip), the surface is relatively smooth. For areas of 

low to moderate sliding, the surface appears to have some 

plastic flow (rippling) and evidence of abrasive wear. Note 

that some wear products remained on the surface after the 

mild mechanical cleaning that was done prior to 

measurements. The visual appearance of the teeth surfaces 

were similar for all tests ranging from 20,000 to 80,000 total 

dithering cycles. 
NASA/TM—2007-215008 6

Figure 6.—The typical condition of the worn gear teeth. 

The wear profiles were characterized by the maximum 

wear depth and by the wear areas, and the data were fit to a 

regression model. The wear depth data and regression results 

are provided in figure 7 and the wear area data and regression 

results are provided in figure 8. The wear rate was 

approximately 1.2 nm maximum wear depth per gear dithering 

cycle. The rate of wear area removal was approximately 

5×10–6 mm2

 per gear dithering cycle. Both the wear depth data 

and wear area data show a linear trend of wear accumulation 

for the full range of test cycles that were investigated. The data 

also shows a significant range of results for a given number of 

cycles. To characterize the data, two-sided point-by-point 

95 percent tolerance intervals at 95 percent statistical 

confidence were calculated. The statistical intervals quantify 

the range of data one could expect for a large population of 

tests conducted in identical fashion. By plotting the sorted 

residuals of the regression models, it was verified that the 

assumption of a normal distribution of the residuals in the 

transformed domain was a reasonable assumption. The 

tolerance intervals are provided on the data plots of figures 7 

and 8. 

Discussion 

The experiments and test results characterized the wear of 

8-pitch spur gears having a dithering type motion, a 1.1 GPa 

(160 ksi) maximum Hertz contact stress at the pitch-line, and 

lubricated with a perfluorinated polyether grade 2 grease. For 

such test conditions the gears operate in the boundary 

lubrication regime, and this produced high wear rates relative 

to the wear rates of high speed, oil lubricated gearing. 

Comparing the rate of wear depth accumulation of the present 

experiments to the experiments of Krantz and Kahraman 

(ref. 10) that made use of gears having the same material and 

geometry, the wear rate for the grease lubricated gears of this 

study was on the order of 600 times greater than for oil 

lubricated gears operating in the mixed elastohydrodynamic 

lubrication regime. If such wear rates for the grease lubricated 

gears would not meet the application requirements, then it 

would be necessary to investigate component redesign, surface 

engineering, an improved lubricant, and/or improved antiwear additives to provide the desired performance. 

A significant finding was to discover that the trend of total 

wear was a linear function of test cycles. For the application of 

this data to validate inspection and reuse criteria, it would be 

reasonable to project continued linear trends of field data so 

long as the maximum wear depths and relative percentage of 

case depth loss were bounded by the data of the present 

investigation. 

number of dithering cycles (103 )

0 20 40 60 80

maximum wear depth (mm)

0.00

0.05

0.10

0.15

0.20

maximum wear depth (inch)

0.000

0.001

0.002

0.003

0.004

0.005

0.006

0.007

data

regression

95% tolerance interval

(at 95% confidence)

Figure 7.—Summary of maximum wear depths as a function of 

number of dithering cycles showing datapoints, regression 

model, and tolerance intervals. 

number of dithering cycles (103 )

0 20 40 60 80

wear area (mm2

) 0.0

0.2

0.4

0.6

0.8

wear area (inch2

)

0.0000

0.0005

0.0010

0.0015 data

regression

95% tolerance interval

(at 95% confidence)

Figure 8.—Summary of wear area (or wear volume per unit 

face width) as a function of number of dithering cycles 

showing datapoints, regression model, and tolerance 

intervals.
NASA/TM—2007-215008 7

Summary 

Discussions in support of space mechanism maintenance 

have highlighted the need for data to better understand the 

wear and behavior of mechanical components operated using 

perfluorinated polyether grease. To quantify wear rates for 

gears lubricated with such grease, a set of spur gear 

experiments were completed. The following results were 

obtained. 

1. The visual appearance of the tested surface differed 

from tip to root, likely depending on the amounts of 

relative sliding and surface geometry details. 

2. The visual appearance of the teeth surfaces were 

similar for all tests ranging from 20,000 to 80,000 

total dithering cycles. 

3. The wear rate was approximately 1.2 nm maximum 

wear depth per gear dithering cycle. 

4. The wear rate for the grease lubricated gears of this 

study was on the order of 600 times greater than 

referenced data for oil lubricated gears operating in 

the mixed elastohydrodynamic lubrication regime. 

5. Both the wear depth data and wear area data show a 

linear trend of wear accumulation for the full range of 

test cycles that were investigated. 

6. The data showed a significant range of wear results 

for a given number of cycles. Statistical intervals 

were calculated to quantify the range of data one 

could expect for a large population of tests conducted 

in identical fashion. 

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Handschuh, R.F. and Townsend, D.P., 1996, “Analysis 

of the Effects of Surface Pitting and Wear on the 

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Gears Lubricated With Grease—Observations from 

Seven Experiments,” Proceedings of the 58th Meeting of 

the Society for Machinery Failure Prevention 

Technology, Virginia Beach, VA, April 2004. 

15. Hahn, G. and Meeker, W., 1991, Statistical Intervals—A 

Guide for Practitioners, John Wiley & Sons, Inc., New 

York. 
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1. REPORT DATE (DD-MM-YYYY)

01-12-2007 

2. REPORT TYPE

Technical Memorandum 

3. DATES COVERED (From - To)

4. TITLE AND SUBTITLE

Wear of Spur Gears Having a Dithering Motion and Lubricated With a Perfluorinated 

Polyether Grease 

5a. CONTRACT NUMBER

5b. GRANT NUMBER

5c. PROGRAM ELEMENT NUMBER

6. AUTHOR(S)

Krantz, Timothy; Oswald, Fred; Handschuh, Robert 

5d. PROJECT NUMBER

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5f. WORK UNIT NUMBER

WBS 877868.02.07.03.01.01 

7. PERFORMING ORGANIZATION NAME(S) AND ADDRESS(ES)

National Aeronautics and Space Administration 

John H. Glenn Research Center at Lewis Field 

Cleveland, Ohio 44135-3191 

8. PERFORMING ORGANIZATION 

 REPORT NUMBER

E-16195 

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National Aeronautics and Space Administration 

Washington, DC 20546-0001 

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NASA, ARL 

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NASA/TM-2007-215008; ARL-TR-4124 

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Unclassified-Unlimited 

Subject Category: 37 

Available electronically at http://gltrs.grc.nasa.gov 

This publication is available from the NASA Center for AeroSpace Information, 301-621-0390

13. SUPPLEMENTARY NOTES

14. ABSTRACT

Gear contact surface wear is one of the important failure modes for gear systems. Dedicated experiments are required to enable precise 

evaluations of gear wear for a particular application. The application of interest for this study required evaluation of wear of gears lubricated 

with a grade 2 perfluorinated polyether grease and having a dithering (rotation reversal) motion. Experiments were conducted using spur 

gears made from AISI 9310 steel. Wear was measured using a profilometer at test intervals encompassing 10,000 to 80,000 cycles of 

dithering motion. The test load level was 1.1 GPa maximum Hertz contact stress at the pitch-line. The trend of total wear as a function of 

test cycles was linear, and the wear depth rate was approximately 1.2 nm maximum wear depth per gear dithering cycle. The observed wear 

rate was about 600 times greater than the wear rate for the same gears operated at high speed and lubricated with oil. 

15. SUBJECT TERMS

Wear; Gears; Greases; Lubrication 

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