ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
Figure 3.2.1  Theoretical and Experimental Conductance vs. Axial Load for a  28-mm OD Dry Static Steel Bearing 3.3 Oil Lubricated Static Non-Run-In Bearing Once oil was introduced into the bearing, thermal conductance increased significantly and became relatively insensitive to axial loads. Figure 3.3.1 compares the conductance of a dry static bearing with the  same bearing after the addition of oil. The small bearing was used for this experiment, after oil was added  according to the method described above. The bearing was not operated to run-in the lubricant. Figure 3.3.1  Experimental Thermal Conductance vs. Axial Load for a  Static 28-mm OD Oil Lubricated No n-Run-In and Dry Steel Bearing The green triangles in Figure 3.3.1 (oiled bearing) show a significant increase in conductance over the  red squares (dry bearing); this is attributed to the increase in ball-to-race contact area due to the presence of the oil between the contacts. We surmise that increased conductance is due primarily to the 0.00 0.01 0.02 0.03 0.04 0.05 0.06 0.07 0.08 0.09 0.10  0 50 100 150 200 250 Conductance (W/C)  Axial Load (N) Analysis  Test  0.000 0.025 0.050 0.075 0.100 0.125 0.150 0.175 0.200  0 50 100 150 Conductance (W/C)  Axial Load (N) Before run-in, static, steel oil lube  Static, dry steel  P=40NP=84N 118
oil meniscus surrounding the contact and secondarily due to the liquid mediated Hertzian contact (filling  the microscopic pores between the rough contacting surfaces). In essence, the oil increases the thermal  pathway, decreasing resistance and increasing conductance. The decreased sensitivity to axial loading can also be attributed to the increased initial area due to the  presence of the oil lubricant. With a much larger initial area, a subsequently greater increased area  change is needed to affect bearing thermal conductance. Figure 3.3.2 shows the relation between total  ball-to-race contact area and conductance. The red dots indicate the increased contact area in a dry  bearing showing the change in conductance as a function of load. The yellow dots show the increased contact area in an oil-lubricated bearing due to the same increase in load, and its influence on  conductance. The change in an oil-lubricated bearing is dramatically less pronounced because the  relevant contact area is already an order of magnitude larger due to the oil. 

Figure 3.3.2 Theoretical Conductance vs. Total Inner Race Contact Area for a  28-mm OD Steel Bearing:  Change in  Area for a Dry and Oil Lubricated Case The exact thermal conductance magnitude depends on the initial quantity and distribution of the lubricant. If we assume that the contact area of the lubricant, A 0, is constant, then, bearing thermal conductance is  proportional to the square root of the total area of the lubricant and the ball to race contact: /g1833/g1503 /g3493A+A /g2868 (3.14) The ball to race contact area, A, is still proportional to the axial load to the 2/3 power:    /g1827/g1503/g1842/g2870//g2871(3.15) Measuring conductance, G, the Yovanovich model can be used to estimate the total area A+A 0 [3], then,  the ball to race contact area can be calculated based on knowledge of the load, P [3]. The assumption is  that the thermal resistance due to the lubricant is negligible, despite the thermal conductivity of oil being  only 0.3% of the 52100 steel. The thermal resistance is still insignificant because the thickness of the  lubricant in the meniscus at rest is, on average, approximately 0.3 microns. This assumption introduces 0 0.02 0.04 0.06 0.08 0.1 0.12 0.14 0.16 0.0E+00 2.0E-02 4.0E-02 6.0E-02 8.0E-02 1.0E-01 1.2E-01 1.4E-01 1.6E-01 1.8E-01 Conductance (W/C)  Inner Race Hertzian Contact Area (mm2)  P=40NP=84NDry Oil P=40NP=84N 119
Table 3.4.1 Static Bearing Ball-to-Race Hertzi an and Lubricant Area After 6 krpm Run-In   Bearing Details Area – (Inner Race) Bearing SizeOD  (mm)Load, P  (N)Hertzian  Contact  Area (mm2)Lubricant  Area  (mm2)Total  Area  (mm2)Lubricant/  Hertzian Area (%) Large bearing, steel 62 133 8.2E-02 11.4E-02 19.6E-02 138 Small bearing, steel 28 129 4.5E-02 8.2E-02 12.7E-02 183 Figure 3.4.1  Comparison of Thermal Conducta nce vs. Axial Load for an Oil Lubricated  Static Before Run-In, After Run-In, and Dr y Conditions for a 28-mm OD Steel Bearing Figure 3.4.2  Comparison of Thermal Conducta nce vs. Axial Load for an Oil Lubricated  Static Before Run-In, After Run-In, and Dr y Conditions for a 62-mm OD Steel Bearing0.00 0.05 0.10 0.15 0.20 0.25 0.30  0 50 100 150 200 250 Conductance (W/C)  Axial Load (N) 28mm OD, Before run-in, static, steel oil lube  28mm OD, After run-in, static, steel oil lube  28mm OD, Static, dry steel  0.00 0.05 0.10 0.15 0.20 0.25 0.30  0 50 100 150 200 250 Conductance (W/C)  Axial Load (N) 62mm OD, Before run-in, static, steel oil lube  62mm OD, After run-in, static, steel oil lube  62mm OD, Static, steel dry  121
larger bearing having much higher conductance than the smaller bearing and rising more steeply with  axial load. Figure 3.5.1 Comparison of Thermal Conductance vs . Axial Load for an Oil Lubricated Static  And Dynamic States of Motion for a 28-mm OD Steel Bearing Figure 3.5.2 Comparison of Thermal Conductance vs . Axial Load for an Oil Lubricated Static  And Dynamic States of Motion for a 62mm OD Steel Bearing0.00 0.10 0.20 0.30 0.40 0.50 0.60  0 50 100 150 200 250 Conductance (W/C)  Axial Load (N) 28mm OD, After run-in, 6k rpm, steel oil lube  28mm OD, Before run-in, static, steel oil  lube  28 mm OD, After run-in, static, steel oil lube  28 mm OD, Static, dry steel  0.00 0.10 0.20 0.30 0.40 0.50 0.60  0 50 100 150 200 250 Conductance (W/C)  Axial Load (N) 62mm OD, After run-in, 6k rpm, steel oil lube  62mm OD, Before run-in, static, steel oil lube  62mm OD, After run-in, static, steel oil lube  62mm OD, Static, dry steel  123
Figure 4.2.1 Run-In Effects on Thermal Conductance of a Virgin Bearing;   Medium Bearing A Figure 4.2.2 Run-In Effects on Thermal Conductance of a Previously Used Bearing;  Medium Bearing B We conclude form this that a bearing of this size, lubricated as these were, will have enough excess oil to  enable a static conductance of 0.16 W/°C. After run-in, a portion of this oil is displaced from the rolling  path, but remains close, while another portion is displaced to a location where it is far from the rolling path. This loss results in a smaller oil meniscus and conductance of 0.10 W/°C. With prolonged rest, the oil that remained close by is able to creep back to the ball-to-race interface, causing a recovery of the  conductance to about 0.13 W/°C.0.060.100.140.180.220.26 0 2000 4000 6000 8000 10000Conductance (W/K) Speed (RPM)Run 1; inc Run 2; inc Run 3; dec Run 4; inc Run 5; dec Run 6; inc Run 7; dec 0.060.100.140.180.220.26 0 2000 4000 6000 8000 10000Conductance (W/K) Speed (RPM)Run 1; inc Run 2; inc Run 3; inc Run 4; inc Run 5; dec 125
In both cases, additional run-in resulted in a decrease in static thermal conductance. However, the  dynamic bearing thermal conductance behavior began quite differently between A and B. For the virgin  bearing A, the thermal conductance began high, but with additional run-in the value decreased. We  suggested earlier that the mechanism of heat transfer in a dynamic bearing was mass transport. The test data indicates that the lubricant film thickness decreases with time then stabilizes after the bearing is completely run-in. That film thickness is established by the centrifugal forces generated by the maximum  run-in speed. The film thickness on the ball should /g81/g82/g87be confused with the Elastrohydrodynamic (EHD)  film as the EHD film is the minimum oil thickness separating the ball and race during operation. Instead,  the film thickness that we are referring to is the lubricant riding on the outer diameter of the ball. The second bearing, B, had previously been run-in and clearly shows that the dynamic conductance  values are repeatable. This suggests that the lubricant film on the ball does not recover after run-in. The testing of both medium bearings varied in the number of start/stop cycles and the operational time,  but both had seen maximum run-in speeds of 10,000 rpm and were operated at the same temperature. The conductance values after multiple speed cycles, shown in Figures 4.2.1 and 4.2.2, are very similar. Both the static and dynamic bearing thermal conductance values ended up comparable, suggesting final run-in lubricant quantity and distribution in the bearing is similar and is governed primarily by maximum  run-in speeds. 4.3 Influence of Maximum Run-In Speed on Lubricant Loss The bearing thermal conductance profile changes, as the bearing is run-in at even higher speeds (20,000  rpm). This effect is shown in Figure 4.3.1. Figure 4.3.1  Effect of Preconditioning Bearing at 10 krpm and 20 krpm Run-in on  Final Bearing Thermal Conductance fo r Medium Sized Steel Bearings  After operating at higher speeds, the thermal conductance decreases indicating that more oil has been  displaced. We surmise that this oil has been driven away predominantly by higher centrifugal forces and  possibly to a lesser extent by higher thermal gradients.  4.4 Influence of Rest Time on Lubricant Redistribution For the final series of tests, one medium bearing (Medium Bearing B) was allowed to rest to observe the  recovery of oil that had been displaced in earlier tests. After the suite of testing shown above, the test fixture remained dormant for three months at room temperature. During this time, the bearings did not 0.080.100.120.140.160.180.200.220.240.26 0 5000 10000 15000 20000 25000Conductance (W/K) Speed (RPM)10 krpm Run-In 20 krpm Run-In 126
spin. However, at three points in time during this interval, the static conductance was measured. Figure  4.4.1 indicates a slow growth in the static conductance. This growth is consistent with the recovery of oil  at the ball/race menisci. We find a slow equilibration process, resulting in a thermal conductance increase of about 45% after  resting for four months. This corresponds to a lubricant area increase of over 160% in the ball-to-race pathway for a bearing at static rest. This suggests a possible strategy for remediation of mechanisms that are suffering from lubricant depletion. Figure 4.4.1 Static Bearing Thermal Conductance vs. Rest Time  For Steel Medium Bearing A    5.0 Summary and Conclusions The lifetime of precision bearings in space applications, where typically only a single charge of lubricant is  expected to provide for the entire mission life, is closely tied to the lifetime of that lubricant. Mechanisms  often function well until the lubricant is depleted, and then rapidly deteriorate. Some lubricant may be lost to thermo-chemical degradation during normal operation. However, in many situations much of it is simply displaced from the critical interfaces. Among this displaced lubricant some may be recovered by surface  creep when the force balance is changed, while some is unrecoverable due to its having been thrown  from the bearing or located with other impediments to migration. It would be helpful to understand the  mechanisms of lubricant loss and recovery.   These experiments have shown that thermal conductance measurements are sensitive to lubricant  quantity. For this reason, thermal conductance measurements can be used as a tool to make  observations on lubricant behavior, such as run-in effects, and impact of rest times on lubricant redistribution. We show that conductance after run-in can be sensitively dependent on bearing geometry  and operating speed, and less dependent on initial lubricant state, in the conditions studied here. These results imply that lubricant distribution is impacted by operational conditions, changing thermal conductance in logical ways. Some of the oil that is displaced by run-in can be recovered during rest,  presumably by surface migration back to the contact, increasing conductance, but this process required  months of time in the case studied here.   0.090 0.100 0.110 0.120 0.130 0.140 0.150  0 20 40 60 80 100 120 140 Conductance (W/C)  Rest Time (days)  127
References  1. Y. R. Takeuchi, S. E. Davis, M. A. Eby, J. K. Fuller, D. L. Taylor, M. J. Rosado, “Bearing Thermal  Conductance Measurement Test Method and Experimental Design,” Rolling Element Bearings: 9th Volume, ASTM STP 1542, Y. R. Takeuchi and W. F. Mandler, Eds., West Conshohocken, PA, 2012. 2. P. A. Bertrand, D. J. Carré, Reinhold Bauer, “Oil Exchange Between Ball Bearings and CottonPhenolic Bail-Bearing Retainers”, Tribology Transactions, 1995; 38:342-352.  DOI:10.1080/10402009508983414  3. Yovanovich, M. M., Analytical and Experimental Investigation on the Thermal Resistance of  Angular Contact Instrument Bearings , Instrumentation Laboratory, E-2215, Massachusetts  Institute of Technology, Cambridge, Massachusetts, Dec. 1967.

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responsible for its failure. The SEM used a cold-field emission electron source with tungsten single-crystal  tip. The sample chamber operated at a vacuum of 4x10-7Torr. To prevent charging, if necessary, a low  accelerating voltage of 1kV was used to image the bearing component, particularly the Si 3N4ball. At such  a low accelerating voltage, it was not necessary to coat the ball being examined with a conducting layer, such as pyrolytic carbon or gold.  Inducing Artificial Scars on Si 3N4 Balls A technique was created to generate scars with dimensions and morphology that are comparable to the  distinctive features observed from contractors and in as-manufactured Si 3N4balls (Figure 2). The  technique involves controlled back-and-forth sliding contact of two Si 3N4balls. This ball-on-ball apparatus  is shown in Figure 3. The apparatus consists of a linear stage to enable precise relative translational motion between the balls, and two ball holders with conforming hemispherical sockets. The balls rest inside the sockets and their rotational movement is constrained by rubber-tip set screws.  Figure 2.  A typical scratch found on Si 3N4ball (left), and a distinctive feature observed by contractors (right). Figure 3. Ball-on-ball test fixture used in generating artificial scars on Si 3N4balls. Because it was necessary to generate scars of the same size and morphology on the Si 3N4balls for each  RCF test campaign, a technique was developed to improve the consistency and reproducibility of the scars. The Si 3N4balls had to be cleaned by soaking in a warm bath of isopropyl alcohol for a few minutes  and drying completely with lint-free cloth. Two balls were then placed in the ball-on-ball apparatus and  locked in place. As a check for cleanliness, a clean swab with a spongy tip was moved across the surface  of the ball. The surface of the ball was deemed clean when it felt like the ball was “grabbing” onto the  moving swab. A final quick cleaning with a cotton/sponge swab removed any dust/lint that might have deposited in the contact zone. An aluminum block of a known mass was placed on top of the fixture to supply a load, which set the amount of material removed from the ball per pass and, ultimately, the final depth of the scar. Several back-and-forth passes were induced. The total travel per pass was 1.27 mm  (0.050 in) for the “large” scars with a width of approximately 100 μm, and 0.64 mm (0.025 in) for the  131
Figure 5. Micron element size FEM (left) and whole model (right) region of interest showing the  high tensile stresses created by the traction load.  Figure 6. Close up views of two parametric mode ls showing the tie command regions with one  significant mesh size transition (left) and two significant mesh size transitions (right).  Validating Finite-Element Models with Three-Dimensional Explicit Equations for Sliding Spherical Contact Additional confidence in the three dimensional stress results from the FEMs was required to predict the  stress distribution and propensity for crac k propagation due to sca rs and cracks on a Si 3N4ball. There  are numerous variables that can influence the validity of a FEM solution. Therefore, the three-dimensional stress field of the finite-element analysis was compared to an explicit closed-form analytical solution. After  conducting a literature search, an appropriate set of equations was identified for sliding spherical contact [7]. This analytical solution could determine the interior stresses as well as the surface stresses of the ball due to sliding contact, differentiating it from the Hertzian solution used in previous FEM validation efforts. Validating the FEM stress results from the ball’s interior was integral to establishing realistic stress fields  for common ball defects. For visualization purposes, the analytical solutions were programmed into a  subroutine that was executed by the finite-element analysis program. Stress results from the analytical solution were calculated at each integration point in the FEMs and saved to the output database. This procedure enabled side-by-side comparisons between the analytical and the numerical results.  Normal and shear stress results from both the explicit solutions and FEMs showed good agreement. The  overall character of each stress measure from both the analytical and numerical results matched, while  the absolute maximum and minimum stresses differed. The FEM results consistently presented higher maximum stresses and lower minimum stresses for each stress measure. Stress results in compressive regions were within five percent of each other, while differences in the tensile regions varied more.   Effect of Scars on the Life of a Si 3N4Ball Modeling efforts supported rolling contact fatigue testing to determine the effect of scars on hybrid ball  bearing life. Scars were modeled in the FEMs by deleting elements in a manner to mimic the geometry of  133
a typical lab generated scar. The scar was placed in the high tensile region of the FEM caused by the  traction load because that was the most probable failure initiation region. A raceway contact simulation was run and the FEM results demonstrated little change in the Si 3N4ball’s stress field around the contact  region due to the presence of the scar (Figure 7). Therefore, a scar was deemed benign in terms of the  effect on the life of the Si 3N4ball.  Figure 7. Maximum principal stress of contact with  a scar (left) and without a scar (right).  However, during experimental testing, early failures in the hybrid ball bearings with scars indicated the  bearing was affected by the presence of scars. Spalls prematurely developed in the steel raceways. When the hybrid bearings were disassembled, the scars on the Si 3N4balls showed signs of wear but did  not demonstrate any catastrophic growth.  Effect of Si 3N4Ball Scars on the Steel Raceway A pristine contact condition produced high von Mises stresses in the steel raceway directly under the  contact patch (Figure 8). This high stress region did not extend to the surface of the raceway. The ball scars changed the shape of the stress field, shifting the high von Mises stress region closer to the surface  while creating a low stress region directly under the scar. In addition, at the scar edge a local stress  concentration appeared in the steel raceway and connected with the high stress region. As future work, elastic-plastic analysis will be used to determine the fatigue life of the raceway based on  the abnormal cyclic loading due to the ball scars. An explicit integration two-dimensional FEM is being developed to resolve the stress field in the Si 3N4ball and raceway when a ball scar is located within the  contact patch. Figure 8. Stress contours in the steel from pristi ne ball contact (left) and contact with a 100-μm wide ball scar (right). 134
Database of Full-Scale RCF Tests Baseline (As-Manufactured) Hybrid Bearings /g20/g19/g20/g43/g3/g37/g72/g68/g85/g76/g81/g74/g86/g3 A summary of the test results for the 101H bearings is presented in Table 1. All six tests have been  completed. All tests were conducted at 400 ksi (2.8 GPa) mean contact stress. Out of the six tests, four tests ended due to fatigue failures and two tests reached 3000 hours without failing. For the four tests that resulted in failures, the spall(s) responsible for the failure occurred either in the ball, the inner (IR), or the  outer (OR) race. One of the tests that reached 3000 hours was allowed to run until an outer race spall formed, which gave a total of five fatigue failures from the original six tests. The bearing that had OR  dents for Test #4 was replaced with a new bearing, and the test (Test #7) was restarted and run until a  ball spall eventually suspended the test.  Table 1.  Test results of 101H bearings at 400 ksi (2.8 GPa) mean stress Test # Test Time (hours) Failure Type 1 383 IR spall 2 3054 dents (run-on) 3 38 ball spall & IR spall 4 359 OR dents, replaced (restarted as Test 7) 5 1728 ball spall 6 3353 OR spall 7 1121 ball spall /g22/g19/g23/g43/g3/g37/g72/g68/g85/g76/g81/g74/g86/g3 A summary of the test results for the 304H bearings is presented in Table 2. These tests were conducted  at 400 ksi (2.8 GPa) mean contact stress. Out of the seven tests, six tests have been completed. The six completed tests all developed fatigue failures, but two of the tests surpassed 4000 hours. The failures  were a combination of ball and raceway spalls. One test is still in progress during the preparation of this manuscript with a total run-time surpassing 8000 hours. Table 2.  Test results of 304H bearings at 400 ksi (2.8 GPa) mean stress Test # Test Time (hours) Failure Type 1 2373 ball spall 2 1876 IR spall 3 4821 IR spall & OR spall 4 515 IR spall 5 4384 ball spall & IR spall 6 64 OR spall 7 8000+ still in progress Hybrid Bearings with Artificially-Scarred Si 3N4 Balls A summary of the test results for the 304H bearings with scarred Si 3N4balls is presented in Table 3. All  tests were conducted at 330 ksi (2.3 GPa) mean contact stress. Fractographic images of failure features are presented in the next section. 135