ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
0.1110100 - 8 0- 6 0- 4 0- 2 0 0 2 0 4 0 6 0 8 0 Temperature (degrees C)Lubricant Film ParameterFull Separation Regime Acceptable EHD Regime Difficult Rock - COBRA EHL (750 rpm) Difficult Rock - Hamrock-Dowson (750 rpm) Difficult Rock - Zaretsky Starvation Adjustment (750 rpm) Test Temperature1.53.0

Figure 3: Lubricant Film Parameter versus Temperature 

The remaining question with regard to the bearing and lubrication design was how much lubricant to use.  To minimize parasitic torque, especially at temperatures less than -45° C where Brayco 815Z viscosity  starts to increase dramatically, grease plating the bearings would be ideal. However, this means the bearings would likely be starved for lubricant and the analysis showed that boundary contact was  extremely likely at the higher end of the temperature ra nge. When considering that local self-heating of  the bearing would elevate temperatures even more, boundary contact seemed practically guaranteed. It  was decided that a verification test must be performed to demonstrate that the bearing design would  survive the 3x-life requirement. Ro om temperature (20° C) was selected as the test temperature for  convenience and because it rep resented a reasonable maximum ambient based on past RAT experience  – very rarely was the RAT operated at ambient temperatures above 20° C. Finally, a decision was made  to test two bearing sets side by side. One bearing set would be grease plated only, while the other set  would be grease plated and filled to 10 vol% with a grease/oil slurry. 

Experimental  Test Bearings

The NHBB bearings (MER-1960SD501 DB) used for the test were procured from a distributor, the Intercontinental Bearing & Supply Company (IBSCO). These bearings had been precision ground for a  preload of 27 N (6 lb). The bearings were also lubricated with a MIL-L-6085C synthetic diester oil  (Winsorlube L-245X). Several steps were necessary to make the bearings suitable for the SRT bearing life test. 

First, the Winsorlube oil needed to be replace d with Brayco 815Z.  IBSCO handled this process – a  process they considered proprietary and certified by NHBB. Because the bearing featured a phenolic  retainer, there was significant concern about the re-lubrication process. The concern stems from the fact  that porous cotton phenolic retainers, if not sufficiently saturated, can absorb oil from the bearing contact  zone [12][13][14]. It is evident from these reports that typical vac uum impregnation processes used in  industry cannot fully saturate cotton-phenolic retainers. The retainers continue to absorb oil over the  course of years until fully saturated. Furthermore, once fully saturated, there is no significant transfer of oil from the cage to a poorly lubricated raceway or ball. Therefore, it is very important to saturate the retainer  with as much oil as possible to minimize the sink effect and prevent a premature starved condition.  Members of the design team were permitted to witness the process and bearing mass measurements  74
taken at various points of the process yielded an estimated average of 19.5 mg of Brayco 815Z oil  absorbed by each retainer+. This represented a >5 wt% increase to the retainer, which satisfies one  spacecraft standard [11], and nearly 85% saturation assuming a 5% porosity.     Second, two bearing sets were grease-plated using a 10/90  wt% grease-plate solution of Braycote  micronic 601EF and Fluoroclean X100. The bearings were delivered with a thin coating of Brayco 815Z  and grease plating with this solution ensured that the 52100 steel components were coated with the grease’s corrosion inhibiting additive (sodium nitrite). Grease plating was accomplished by pouring  enough Fluoroclean X100 into a clean glass jar to fully cover a bearing. The Fluo roclean mass was  recorded to allow for a calculation of the amount of Braycote Micronic 601EF to add. Once the Braycote micronic 601EF was added and mixed into solution, the grease-plate solution was ready for use. Each  bearing was dipped in the solution for 5-10 seconds, placed on a Class 100 clean bench for 10-20  minutes; followed by an oven bake at 66/g219C for 60 minutes. Once the bearings were cooled, they were  stored until needed.     Third, one of the bearing sets was filled an additional 10 vol% with a 50/50 wt% mixture of Braycote  micronic 601EF and Brayco 815Z oil. The 50/50 mixture had a lower viscosity than pure grease allowing  for cold-temperature ope ration on Mars. A CAD model was used to  estimate the empty space inside the  bearing to allow for determination of the correct mass of the 50/50 slurry to add. A total of 170 mg of the  slurry was added to bearing pair using a syringe.   Finally, in order to achieve the desired preload, a shim was added between the outer races of each  bearing pair. To determine the proper shim thickness, measurements were made on the bearings to verify  the preload deflection of the bearing pair and correlate COBRA software preload predictions.      Motor Radial Load  Gage Load Application  Housing Axial Load Gage  Torque Sensor  Housing  Figure 4: Bearing Test Stand  

                                                       + Subsequent analysis performed by NASA JPL showed that some amount of the Winsorlube remained in  the retainer post-processing, however the analysis procedure requires refining to determine the exact  amount.  75
Test Apparatus   The bearing test stand (Figure 5), capable of simultaneously testing two bearing configurations, consisted  of housing and shaft components for mounting and loading the test bearings, mechanical force gages,  brushless DC drive motors, and torque sensors. Each test bearing set was mounted on a “non-rotating” shaft (shown in Figure 5) which was rotationally grounded through its torque sensor – in this way, the  drag torque in the bearing could be directly measured. To achieve the desired combined bearing load  case, independent axial and radial loads were applied to the non-rotating shaft by adjusting load screws with in-line dial force gages. The position at which the radial load was applied could be adjusted in order  to achieve the desired moment. A tapered roller bearing between the load application housing and the  non-rotating bearing shaft attempted to rotationally isolate the shaft from the loading gages so that bearing drag torque is reacted only by the torque sensor. 

Test BearingsTorque  SensorMotorRadial Load  Gauge Axial  Load  GaugeLoad  Application  Housing Non-rotating  Bearing Shaft

Figure 5: Bearing Test Stand Cross-Section 

The entire test stand was placed in a small vacuum chamber with gas feed-thrus and an automatic  control system. The controller maintained the chamber pressure at between 0.27-1.6 kPa (2-12 torr) and  with each pump cycle replenished the CO 2 atmosphere by injecting a small amount of the gas.      

Test Procedure   Testing was broken up into six blocks; each block representing one-half of the total required bearing life. At the conclusion of the test, three times the required life would be demonstrated. Each block was broken  up into four load cases (shown in Table 5) representing expected operating conditions. This methodology  allowed the various load cases to be spread across the entire life. Conducting each load case test run consisted of the following steps:  /g120 Run 1 st pre-test diagnostic to verify that test fixture was functioning properly;  /g120 Adjust applied bearing shaft loads & configure test fixture in vacuum chamber;  /g120 Run 2nd pre-test diagnostic to verify that loade d system was functioning properly;  /g120 Purge (CO 2) and evacuate chamber to 0.27 kPa (2 torr);  /g120 Run test at case loads and spe ed for the specified amount of time;  /g120 Break chamber, unload bearing shaft and run post-test diagnostic to confirm that test fixture was  still functioning properly. 

76
Figure 6: Bearing Life Test Data 

010203040 0.0E+00 1.0E+07 2.0E+07 3.0E+07 4.0E+07 5.0E+07 6.0E+07 Bearing RevolutionsTorque Noise (mNm)10/90 Grease Plate (Min/Max delta) Noticable shiftSteadil y  increasing BaselineTorque limit  exceeded fo r  1st timeCatastrophic  failure

Figure 7: Torque Noise in Grease-Plated Bearing (BP-A) 

Following the test, the fixture was disassembled an d it was noted that BP-A was no longer preloaded  even though the preload fastener was still tight – a gap of 0.64 mm was measured between the outer races. Upon removing the bearings from the fixture, the races separated with very little force – the  bearing are non-separable by design. A substantial amount of brownish-black debris was present on BPA and many of the fixture components near the torque sensor (see Figure 8). It was determined that this  debris was generated by frictional wear between the stainless steel torque sensor coupling (Figure 8,  right image) and the aluminum cylinder that housed the outer races of BP-A. The races of BP-A had worn so much that the non-rotating bearing shaft was permitted to cock several degrees. This allowed the  78
components to rub together generating debris that contaminated the bearings. Without being able to  pinpoint exactly when this began, it’s likely based on the torque signature that the race wear became  severe enough to allow rubbing some time during Lo ad Case 2 of Block 5 leading to the first torque overlimit event. 

Coupling  with debrisBP-A Non-rotating  shaft

Figure 8: Post-Test BP-A (Left – hand separable bearing; Right – shaft with particulate)   Further BP-A lubricant observations were not possible due to the excessive contamination by wear  particulate shown in Figure8. However, further inspection of the bearing components was conducted using a stereomicroscope.  Prior to this inspection, BP-A components were soaked in methylene chloride  for 90 minutes. The solution was agitated manually to aid in removing any particulate captured within the  bearing cavities. Following the methylene chloride soak, the bearing components were cleaned with isopropyl alcohol and the metallic components were run through a 3-stage ultrasonic cleaning procedure.   Microscopic inspection showed that noticeable wear tracks were present on both the inner and outer  races (see Figure 9). Very heavy wear existed at two distinct circumferential locations on the inner race  while there was more evenly distributed wear on the outer race. This was to be expected since the outer  race was rotating and the inner race was stationary with respect to the applied load. With the possible exception of very minor pitting, the balls appeared to be in decent shape. The phenolic retainer also did  not exhibit significant wear. 

Outer Race Inner Race Wear  Tracks

Figure 9: BP-A Components Post-Test (clean) 

Bearing Pair ‘B’ (BP-B) – grease-plate plus 10 vol% grease fill   Through 3x life (5.96 x 107 revs), the average torque remained at approximately 5 mNm or less –  however, it is suspected that the tapered bearing affected the torque sensor signal at the higher load  cases (3 and 4). BP-B also ran much quieter than BP-A for the totality of the test with the torque noise averaging between 1-2 mNm. At 2.6x life (5.25 x 10 7 revs), the first hints of wear started to appear in the  torque noise signal with the noise spiking no higher than 4 mNm. 

During disassembly of the fixture to remove BP-B after the test, no noticeable axial or angular play was  found in the test shaft. There was no gap between the outer races of the duplexed pair. When rotated by hand, the bearing pair ran smoothly. A visual inspection confirmed that there was still an oil-grease slurry  in the bearing (Figure 10). The bearing pair was weighed before and after the test to determine the  amount of lubricant lost. The total mass loss was 56 mg representing 33% of the original 170 mg of slurry  79
added to the bearing before the test. This loss could be due to either the migration of oil from the bearing  due to surface wetting or inertial forces or could represent a mass decrea se due to the catalytic  degradation of the PFPE lubricant. A combination of both loss mechanisms is likely. 

Figure 10: BP-B Bearings Post-Test (Left – Grease-fill; Right – Flouroclean X100 soak) 

A single bearing from BP-B was soa ked in Fluoroclean X100 for ~20 hours to remove the remaining  lubricant and any wear debris internal to the bearing.  Initially, a small amount of translucent brownish  particulate came out of the bearing (Figure 10, right). After 20 hours, most of the particulate had broken up and dissolved. The brownish particulate (sometimes likened to “brown sugar”) is a telltale sign that the  PFPE lubricant was breaking down as a result of the catalytic chemical reaction [16].     Once clean, a slit was cut into the outer race to allow for disassembly. Photographs of the bearing  components were taken and are shown in Figure 11. Evenly distributed wear tracks can be seen on both  the inner and outer races. The inner race does show slightly heavier wear signs at a given location in a  similar manner to BP-A components; however of a much lesser degree. As with BP-A, the balls and  retainer did not exhibit any significant wear.      Outer RaceInner Race Wear  Tracks 

Figure 11: BP-B Components Post-Test (clean) 

As shown above, bearing pair ‘B’ (10 vol% grease-fill) ran with a lower parasitic torque and did not display any severe signs of boundary contact type wear other than some degradation of the lubricant (brown  sugar). This is due to the lack of a starved condition which kept /g513 > 1.5 and allowed operation in the EHD  regime. Bearing pair ‘A’ (grease-plate only) shows intermittent increases in parasitic torque anywhere from two to six times nominal operating torque. These  torque anomalies are followe d by a sudden drop  back down to nominal torque values. As the test progressed, these anomalies increase in frequency and  severity. A similar torque signature has been documented previously and was characterized by significant wear patterns at the ball-cage and raceway-cage interfaces [6]. Also noted in [6], is that the cause of this  type of failure is oil starvation and injection of additional oil during these events caused a decrease in  parasitic torque to nominal levels.    80
Mechanisms and Tribology Symposium, ESTEC, Noordwijk, The Netherlands, 28-30 October 1992  (ESA SP-334, April 1993) 

[8] Nishimura, M. and M. Suzuki, “Tribological characteristics of solid-lubricated ball bearings operated  for 10,000 hours in a vacuum”, Japanese Journal of Tribolo gy, 43(3) (1998) 361-374. 

[9] Loewenthal, S. H., R.G. Chou, G. B. Hopple, and W. L. Winger, “Evaluation of Ion-Sputtered  Molybdenum Disulfide Bearings for Spa cecraft Gimbals”, Tribology Trans., 37(3) (1994) 505-515. 

[10] am_Braycote_601_EF_Mic_USA_English_pds.pdf, <http://www.castrol.com>   [11] “The oil impregnation of phenolic resin based materials used in the fabrication of ball-bearing cages”,  ESA PSS-01-712 

[12] Bertrand, P. A. “Coray 100 Oil Absorption into Cotton-Phenolic Ball-Bearing Retainers”, Aerospace  Corp., El Segundo, CA, Report No. TOR-0091(9975)-4 

[13] Bertrand, P. A. “Oil absorpt ion into cotton-phenolic material”, Materials Research Society Vol. 8 No. 7,  Jul. 1993, Aerospace Corp., El Segundo, CA. 

[14] Bertrand, Carre and Bauer, Aerospace Corp, TR-94, “Oil Exchange Between Ball Bearings and  Cotton-Phenolic Ball-Bearing Retainers” 

[15] Nishimura, M. and M. Suzuki, “Solid-lubricated ball bearings for use in a vacuum”, Tribology  International 32 (19 99) 637-647. 

[16] Space Mechanisms Lessons Learned Study. 07 Jan. 2008. NASA. 18 Dec. 2007.   <http://www.lerc.nasa.gov/WWW/space mech/vol1.html>  82
The bushing test hardware was designed to mimic a portion of the MSL Rover’s Main Differential Pivot  (MDP). The flight pivot assembly is comprised of two flanged bushings located between a central pivot shaft and a hollow cylindrical housing (Figure 2). In the test assembly (Figures 1 & 3) the central pivot shaft was attached to a fixture comprised of an aluminum block, several C-channel sections and a baseplate. The cylindrical housing was attached to a driveshaft driven by a motion controlled DC servomotor through several flexible shaft couplings and a torquemeter. This torquemeter was used to measure the friction torque generated in the pivot assembly during testing.                       Figure 1:  Test Assembly                                             Figure 2:   Flight MDP Assembly  PIVOT ASSEMBLY SERVOMOTOR ASSEMBLYTEMPERATURE CHAMBER  TOR QUEMETER LOADCELL Figure 3:  Bushing Test Hardware in Temperature Chamber  85
Test Temperature / Cycling Profile  The bushing test temperature profile was generated to provide friction torque data throughout the  expected pivot operational temperature range with an emphasis on the low side of range. The low  temperature bias was selected as overall bushing clearance variances were most extreme at the low end  of the temperature range. Also, low temperature friction and wear data for the bonded films and polymeric materials tested was very limited. The bushing test temperature profile and bushing test temperature/dither summary are listed in Table 3.  Table 3:  Phase Two Bushing Testing Dither Cycle / Temperature Profile  Number of Dither Cycles Temperature Comments 10 +23°C Initial check out, chamber door open  200 +23°C Chamber door closed, N2 purge on  30 +47°C 30 +70°C 30 +47°C 30 +23°C 30 -9°C 30 -43°C 30 -73°C 30 -105°C 20 -120°C Cycles added to prevent ice formation on feedthru  10 -130°C Cycles added to prevent ice formation on feedthru  4720 -135°C 30 -105°C 30 -73°C 30 -41°C 30 -9°C 200 +23°C Dither Cycling Profile  The dithering motion profile run during the bushing testing was formulated based on the output of Matlab  code written by Jaime Waydo [2]. Her Dither Analysis Code  calculated the main pivot (differential) dither  angles produced by rock encounters during a simulated 60-kilometer vehicle traverse in a 20% rock field. The code output dither angle amplitude events for the 60-km traverse as shown in Table 4.  Table 4:  Dither Event Count for 60 km Vehicle Traverse in 20% Rock Field  Main Pivot Dither Angle ( /g84) Range Number of events for 60 km traverse % of Total  /g84 < 5 degrees 3252 78% 5 degrees < /g84 < 15 degrees 450 11% 15 deg < /g84 < 20+ degrees 439 11% Based on this output a simplified test dither cycle profile was formulated. The profile was based on three  distinct dither amplitude cycles, a small angle cycle (±2 degrees), a medium angle cycle (±10 degrees)  and a large angle cycle (±20 degrees). When these groupings were mapped to the frequencies generated by the Matlab code the following profile was developed. A 10-cycle set started with a ±2087
degree dither cycle followed by four ±2-degrees dither cycles, one ±10-degree dither cycle, and finally  four more ±2-degrees dither cycles. The final test dither cycle profile is plotted in Figure 4  Figure 4:  Bushing Test Dither Cycling Profile  This dither cycle profile was repeated until the total number of cycles planned at each temperature target  was reached and the drive assembly was paused. The total number of dither cycles run for each test was based on a 60 kilometer total scaled to 80 kilometers (or four times the current 20-km mission traverse distance requirement). This scaling yielded a total bushing test dither cycle number of 5520 (for a 4x test). The maximum angular rate of the servomotor output was set to 2.5 RPM for the duration of the dither  testing. This angular rate was selected based on a combination of actuator/torque capabilities and expected maximum differential articulation rates during traverse. Bushing Loading  The radial loads placed on the test bushings were intended to replicate a subset of those imparted on the  flight MDP bushings during the primary mission of the vehicle. Dynamic loading of the bushings due to the touchdown or multi-wheel drops during traverse was not factored into the load calculations for the test bushings. Only the six wheels flat, 1G, 30 deg vehicle roll Adams model output values scaled to 3/8g equivalent loading were used to determine the target loading for bushing testing. The flight bushing radial loads and test loads are listed in Table 5. Also listed is the average bearing stress on the bushings due to radial loading and approximate contact stress values based on formulas from Roark’s Formulas for  Stress and Strain [3]. Table 5:  Bushing Loading Summary  Bushing Location Radial Bushing Load (N) Average Bearing  Stress (kPa) Contact Stress (kPa)  MDP Inboard Bushing 5383 3.6 (0.520 ksi) 14.5 (2.1 ksi)  Test Inboard Bushing 6427 4.3 (0.622 ksi) 15.9 (2.3 ksi)  MDP Outboard Bushing 5127 4.9 (0.708 ksi) 17.9 (2.6 ksi)  Test Outboard Bushing 5267 5.0 (0.727 ksi) 18.6 (2.7 ksi)  88
The loading method employed in the bushing test setup did not allow the applied radial load to be  matched for both the inboard and outboard bushing. As such the loading in the inboard test bushing was slightly higher than the target flight MDP inboard bushing radial load. Although the moment-carrying load path for the flanged bushing pair was assumed to be purely radial for  the loading calculations listed in Table 5, actual bushing loading in the test assembly proved that this  assumption was at least partly flawed. The most significant bushing wear in the initial tests was on the  edge of the thrust face of the bushing indicating that at least part of the moment applied to the test setup  was being reacted out on the thrust faces of the bushings. This result was not altogether unexpected as  the test pivot assembly geometry does have a redundant moment carrying capability through the thrust faces of the bushings. The radial gaps in the pivot required due to dissimilar coefficient of thermal  expansion values between the gall resistant stainless steel bushing material and the pivot structure allowed this redundant load path to be possible. Although the loading produced in test setup is still thought to be representative of the flight environment, exact contact stress value determination for the  bushings was significantly complicated by this non-ideal loading. If this tribological combination is  selected for future designs careful thought should be employed to produce at design with a more tightly constrained load path.  Results Friction Torque The friction torque measurements obtained during the bushing test series are plotted in Figures 5 through  11. Housing temperature (shown in red) during the measurements was also plotted on a secondary axis  for reference. Each plot was generated with the same scale on the X and Y axes to permit a quick visual comparison of the friction torque profile for each test.  Figure 5:  Friction Torque Plot  “Phosphate-like” Binder MoS 2 Bonded Film Coated Bushings vs. Anodized Aluminum)  89
Silicate binder MoS 2 bonded film coated Stainless              Phosphate-like binder MoS 2 bonded film coated Stainless Figure 13:  Post-Test Photos of Smaller, Outboard Bushing                    Silicate binder MoS 2 bonded film Test                                      Phosphate-like binder MoS 2 bonded film Test Figure 14:  Post-Test Photos of Wear on Aluminum Pivot End Plates                                   Polyamide-imide Test                                                                           Polyimide Test Wear scar area,  steel exposed Wear scar area, steel partially  ex posed 95