ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
308  

Figure 5: Boundary  Lubrication Wear Rate at 75°C  

In atmospheric conditions, the wear rate for 52100 steel was consistently lower than the 440C although  they trended in opposite directions with the increase in lead content.  Under vacuum conditions, the effect  for the sample with 0% lead was minimal but the samples with 3% and 5% lead produced considerably  lower wear rates for both 440C and 52100. This opposite trend between  the 52100 and 440C is believed  to be related to the way lead naphthenate interacts with the chemical composition of the steel [10 ]. With  52100, the lead naphthenate reacts with the iron oxide present at the surface layer and created with wear  debris. When under vacuum there is less iron oxide formation which a llows the lead naphthenate to provide  more wear protection. Regarding the 440C, the lead naphthenate will chemisorb into the chromium layers  of the stainless and provide better wear protection as the concentration increases and the environment  goes from at mospheric to vacuum.  

Comparing the wear performance between 20°C and 75°C, all samples  had a higher wear rate  (2-3X) on  440C at 75°C  except for the 5% lead naphthenate sample tested under vacuum. As this SRV testing is a  pure sliding test, the lower wear  resistance for the 440C at 75°C is tied to the more complex layered  structure of the metal. The structure of the 440C would require both chemisorption and physical absorption  for the best performance. On the 52100 specimens, the samples with 0% lead had h alf the wear at 75°C,  around half the level of wear at a 3% l and 5% loading  of lead under the atmospheric tests.  In the vacuum  tests on 52100, all of the wear rates were comparable.  

MTM (Rolling/Sliding Contra- Rotation)   The results in Figures 6 -7 are for the MAC  fluid with 0%, 3%, and 5% lead naphthenate as well as TPP and  TCP when tested in a mixed and boundary lubricating regime and under a rolling/sliding configuration on  the MTM at 50°C and at 150°C. These plots illustrate  the wear  rate of each material.
309   Figure 6: Mixed Film Scuffing Wear Rate Comparison  

Figure 7: Boundary Scuffing Wear Rate Comparison  

From these results, we can see that there is a transition point in the concentration of lead in these MAC  fluids and how it react s with the surface metal to form an anti -wear tribofilm. It should be noted, that this  phenomenon was not seen in the SRV testi ng. It is believed that the aggressive sliding in the SRV test  creates an entirely separate wear mechanism that prevents the lead naphthenate from reacting with the  surface and building a strong tribofilm. In the MTM testing, there is a mixed rolling/slidi ng which will promote  a tribofilm to be created in a fashion similar to gears  and bearings that have a proper run- in process . This  would agree and confirms work done by Carre et al  [11] where it was found in ball bearing test data that  lead naphthenate reacts with metal wear particles to create lead- containing surface coatings.
310 Spiral Orbit Tribometry   The results in Figure 8  are for the MAC with 0%, 3%, and 5% lead naphthenate as  well as TPP and TCP  when tested in a pure rolling SOT test under boundary lubrication conditions. The Rheolube 2000F is a  polytetrafluoroethylene (PTFE)  thickened version of the MAC with 0% lead. In these SOT tests, the TPP  outperformed the TCP and lead based additives in relative life. As the SOT does not provide adequate  conditions to form tribofilms on the surface of the contacts in the test, the TPP is expected to have performed  well due to the decomposition of TPP which creates a multilayered solid film on iron or iron oxide [ 14]. The  results for the TCP and the formulations containing lead were slightly surprising but this is expected to tie  to the restrictions of the mechanics in this test to  create tribofilms and fully simulate a bearings performance.  In all of the SOT testing, the Synthetic Oil 2001A (or 2000F grease version) that contained no additives  performed the best which is believed to be a combination of the structure of the MAC oil itself, inability for  wear additives to create a tribofilms under these test conditions, degradation of the anti -wear additives, and  viscous friction.  

Figure 8: SOT Normalized Lifetimes  

High Vacuum Bearing Testing   The results in Figure 9-10 are for the MAC with 0%, 3%, and 5% lead naphthenate as well as TPP and TCP  when tested in angular contact ball bearings  under boundary lubrication conditions. The grease versions  were also tested with Rheolube 2004 containing 3%  lead. All of the greases  were thickened with sodium  soap.  In these bearing tests, the TPP outperformed the TCP similar to the relative life testing performed on  the SOT . In bearing tests on both the oil samples and grease versions, the addition of lead naphthenate  increased the wear in the bearing which will shorten the lifespan. These results correlate to what was seen  in the SOT testing where the lead was not an effective anti -wear additive for these rolling contacts in  vacuum.
311   Figure 9: Vacuum Bearing Wear Test on Oils  

Figure 10: Vacuum Bearing Wear Test on Greases  

In these  vacuum bearing tests, the Synthetic Oil 2001A that contained no additives performed the best  which is believed to be a combination of the structure of the MAC oil itself, inability for wear additives to  create a tribofilms under these test conditions  (no mechani cal run -in or chemisorption before the test ),  degradation of the anti -wear additives, and viscous friction.   

Conclusions  

In previous studies done on lead naphthenate using the SRV and other sliding tribometers, results have  typically been inconclusive regarding the efficacy of this additive to protect a lubricated contact. The   standard SRV testing includes conditions that could be inappropriate for evaluating many anti -wear  additives including lead naphthenate. This includes the sliding speed which would create a thicker film  preventing the additive function from being studied. The contact stress is also much higher in standard  ASTM tests which will influence how the tribofilms are created and in the case of lead naphthenate, the
316 days. When the calibration sequence is executed, the Secondary Mirror simply rotates to a hard stop at the  WCT position and then returns  to a hard stop at its nominal science position. During this process, the  Secondary Mirror sweeps past the Space View Mirror, thus collecting the information required to correct for  drift in the microwave radiometer signal. The stroke required to rotate fr om the science position to the WCT  position (hard stop to hard stop) is approximately 120° .   Driving Requirements   The primary requirements that  drove the design of the SCS  bearing and housing materials  and arrangement are shown i n  Table 1. Preliminary Design  Review  (PDR ) and Critical  Design Review ( CDR)  values  are shown to highlight changes  in requirement values as the  development progressed. Each  of these requirements had  impa cts on the system design for  different reasons. Program and  design choices made to meet  these requirements ultimately  resulted in the production of this  paper.   Table 1. Driving requirements for the SCS bearing assembly   Requirement  PDR Value  CDR Value  Implication   Schedule  February 2017  July 2017  Relatively short  development time   Mass  < 12.4 kg  < 15.4 kg  All Aluminum  structure   Mass of High Melting  Point Materials  (Titanium)  Not Fully Defined  < 1200 g  Aluminum structure  preferred   Qual Temperature  Range (Op and NonOp the same)  -45°C to +55°C  -45°C to +70°C  Routine  engineering design   Mirror Positioning over  Temperature Range  <0.2/0.2/0.2 mm in each  axis and 0.05/0.05/0.08  degrees including all  launch and in- flight  effects.  <0.2/0.2/0.2 mm in each  axis and 0.05/0.05/0.08  degrees  including all  launch and in- flight  effects . Press fits required  on both inner and  outer races   First Frequency Mode  > 100 Hz  > 100 Hz  Mass Sensitivity.  More mass on  mechanism  requires more  mass on structure.   In the case of mass, the PDR requirement of <12.4 kg was late to flow down which subsequently resulted  in a system well over mass  allocation. The resultant mass at PDR was 14.2 kg Current Best Estimate (CBE)  with 16.3 kg i ncluding Mass Growth Allowance (MGA). At PDR, it was apparent that the 12.4 kg requirement  was going to be a significant challenge to meet going into CDR. A compromise was achieved by increasing  the mass allocation to 15.4 kg, and the design was scrubbed t o enter the CDR with a CBE mass of 12.8   Figure 2. SCS Components
318 reflector.  In the final SCS  configuration, the input assembly simply supported a pinion gear that was mated  with a gear on the output assembly. Gears were not included in the B rassboard assemblies.  

Figure 4. Brassboard test assembly.   The test plan for the B rassboard was designed to measure the sensitivity of the design to the thermal  environment, including thermal gradients  across the bearing races . Multiple tests were performed at room  temperature and qualification hot and cold temperatures to measure bearing drag ( Figure 6). Additional   data were  collected during one of the thermal ramps where a relatively large  (~12°C) gradient was enforced   across the bearing races . For these analyses, ORBIS bearing analysis software  [3] was used to predict  bearing stresses and drag over the temperature range. Overall, results were very promising as predictions  were within 0.1 Nm for most cases.  This torque corresponds to less than 0.1 mNm at the motor which puts  the value wit hin the noise of the system.  Viscosity for the Braycote lubricant was estimated using supplier  data and the model defined by ASTM- D341 to interpolate over temperature.  

Figure 5. Predicted velocity as a function of temperature for two different gearbox efficiencies   Development and testing of the Brassboard provided a lot of value going into CDR. Test results  corroborated bearing drag predictions to within reason and the bearing assembly was proven to perform  within s pecification over the temperature range with no anomalous behavior. F rom an assembly standpoint,
319 there was value in validating and practicing cleaning, lubrication, and assembly procedures. It really  provided a lot of confidence in the performance of the B rassboard design.  

Figure 6. Bearing drag as measured and as predicted over as -built and MMC conditions for the Input and  Output Bearing assemblies  

Instrumenting and Testing the SCS EM in TVAC   Following CDR, a configuration change was made to the output bearing assembly.  There was an  incompatibility  identified in the Brassboard design between the fit required to allow the radial bearing to  float and the mirror pointing requirement. Analysis of the Brassboard configuration pointing requirements  was not explicit ly validated with the Brassboard build. There was no physical space for classic solutions to  this dilemma, such as mounting the radial  bearing in a diaphragm flexure. A floating radial bearing led to a  concern that moment loading of the duplex angular contact pair would be higher than expected during the  launch environment. With this, the team (Honeybee and JPL) agreed to dispense the radial bearing and  change to a more widely -spaced duplex pair,  consistent w ith the input bearing assembly ’s configuration  (Figure 8). The change enabled a more predictable fit through the thermal environment, and therefore a  more predictable constraint on the position of the secondary mirror that the assembly supports. It also  enabled less interference fit on the angular contact bearings, because the clearance fit budget was no  longer being consumed by the radial bearing. The reduction in interf erence fit subsequently reduced  predicted stresses on the bearing pair  while still meeting pointing requirements, as later verified by test.   The SDDA is capable of so much torque, offering abundant  torque margin, that no observable increase  in  SDDA motor current could be measured when operating the SCS spin mechanism. This led to a  configuration change between vibration testing on the EM and Thermal Vacuum testing ( TVAC ). To drive  the Secondary Mirror, the motor and torque transducer combination used to measure drag in the bearing  Brassboard were reconfigured to fit within the SCS assembly. This change produced valuable , high - resolution drag torque data. Actuation of limit switches (used to confirm end of travel and positions around
320 the center of travel)  was easily observed with the torque transducer.  Operation of the SCS through the  torque transducer ended up being key to identifying anomalous behavior in the mechanism. It was also an  important tool in the subsequent anomaly investigation.   Torque transducers were also used upon completion of each bearing assembly to perform a “run- in” test of  the bearings. The purpose was to have the bearing assemblies operating at a “steady -state” torque, for a  given temperature, over the life of the mission. These baseline tests provided a means for evaluating  performance and identifying anomalies in the assemblies throughout their qualification cycle.   Prior to running the first thermal cycle,  a baseline set of data was collected  at +20 °C  (Figure 7). The next  step was  to be actuated at +70°C. However, there was an interrupt in one of the feedback channels used  to control the Ground Support Equipment (GSE) motor. When the SCS returned back to room temperature,  repairs were made to the EGSE  and another set of data was co llected.  Though it wasn’t identified  immediately, there was a ripple on a 12-degree period that became apparent  at +20 °C . This ripple, in  hindsi ght existed following vibration testing, but was amplified after the first ½ TVAC cycle to +70°C. During   the subsequent test at - 45°C, this ripple was impossible to miss.  

Figure 7. EM unit torque profile prior to TVAC  (blue), at room temperature after ½ cycle to +70°C (green),  and at -45°C (yellow)  

EM and Flight Model (FM) Design and Analysis   To discover the root cause of the anomaly, the test program came to a halt. These test results were far  from expected. The Brassboard assembly had shown that there “should” have been no reason for concern  going into TVAC. It validated both performance and analyses. So,  what happened? Something was missed,  but what?
321   Figure 8. Cross -section view showing bearing mount geometry   Bearing analysis was performed using Orbis [1] . For all load cases, predicted contact stresses were lower  than the applicable operating and non- operating limits per GSFC -STD-7000A for 440C steel bearings (335  ksi (2310 MPa) and 400 ksi  (2760 MPa) , respectively , for non- quiet applications ). Alignment across the  temperature range was achieved by controlling the maximum interference  between the bearing races and  mounting surfaces (shaft and housing). Unloaded balls were predicted for the launch load cases at the 3sigma level but none were predicted at the 1- sigma level. Areas of concern were identified at instrument  CDR and closed out with additional analysis. There is still uncertainty related to these aspects of the design,  which are not easily modeled. None have been ruled out as contributing factors to the torque spikes  observed during thermal testing. Neither has any been singled out as the single root cause.   The driving load case for non- operational contact stress is launch vibration. Static equivalent loads were  generated for design purposes for 3 cases: Random X, Random Y and Random Z (all 3 sigma). A single  enveloping load case was constructed using the maximum components and a factor of safety of 1.3 was  applied to the l oad ( Table 2). A 1-σ load case was also constructed.   Table 2. Enveloping static equivalent load case for launch vibration (3 σ and 1 σ ), including 1.3 FoS   Case  Fx (lbf)  Fy (lbf)  Fz (lbf)  Fyy (in -lbf) Fzz (in -lbf)  3 sigma  397 539 390 1755 1246   1 sigma  132 180 130 585 415  Bearing analys es presented at CDR predicted healthy load margins (~ 0.9) against contact stress allowables  when subjected to the 3- sigma enveloping load case.  This is consistent with static load ratings from the  manufacturer ( Table 3). Unloaded balls were indicated  but no truncation. No unloading was indicated for  the 1 σ enveloping load case. Contact stress maps for the 3- σand 1- σ load cases are presented in Fig. 9.  Table 3. Bearing Load Rat ings from Manufacturer   Silverthin  Part #  Physical Specifications  Static Load  Ratings  Dynamic  Load Ratings   Bore  Dia. Outside  Dia. Pitch  Dia. Width  Ball  Dia. No.  Balls  Contact  Angle  Radial  Axial  Radial  Axial   in in in in in deg lbf lbf lbf lbf  SSB040BU5Z  4.0 4.625  4.313 0.313 0.156 58 30 3310  5890  1400  2100
323 thermal cycling or both. Damage caused during vibration could conceivably appear and disappear as the  bearing is thermally cycled and rotated since the balls could track differently under different test conditions.  

Table 4. Random vibration sensitivity analysis results  

Temperature   The driving requirement for selecting bearing fits was the alignment accuracy for the Secondary Reflector  Mirror shown in Table 5. The cantilevered design amplifies the misalignment due to gapping at the bearing.  The minimum press fit for the bearing assembly (largest gap at temperature) and resulting misalignment is  shown.  

Table 5. SCS Secondary Reflector Mirror orientation harmonic error requirement and analytical  predictions presented at CDR  

The effect of temperature on the bearing fits was studied extensively during the anomaly investigation and  it was found that maximum contact stress reported at CDR was underestimated  because the analysis did  not take into account the thicker mount cross -section at one of the bearings and the effect of a steel gear  mounted to the hub (the hub is the outer race housing) . Orbis can account for varying mount cross -sections  but cannot model a bearing mount consisting of two materials with different properties  (i.e. aluminum  housing on bearing row 1 and a steel housing on bearing row 2) . Pressing the bearings into a steel hub  would result in higher contact stresses than pressing into an aluminum hub of the same dimensions.   The results of re- running the analysis w ith these corrections are shown in Figure 10. Due to the CTE  mismatch between the shaft and hub, the trend of increasing stress  versus temperature is reversed. With  an aluminum shaft and hub, the hub grows away from the bearing and the s haft grows toward the bearing  as the assembly is heated. The steel gear, however, restrains the hub radially and the bearing is, therefore,  squeezed at elevated temperature. The highest contact stress is 283 kpsi (1950 MPa) , predicted for a steel  hub that is 20 degrees colder than an aluminum shaft at 70 °C. This stress is significantly higher than
324 temperature- induced operational stress reported at CDR but comparable to the random vibration case 11a  in Table 4 with zero preload.  

Figure 10. Effect of temperature on max mean Hertzian contact stress in the bearing assembly (bearing  installed at 20C)  

The alignment analysis was also corrected for the effects of the thicker sections and steel gear.  At cold  temperature, the misalignment would be increased because the steel gear would restrain the hub from  compressing the outer race. The resulting misalignment predictions , even at a worst case,  are shown in  Table 6. The requirement is satisfied in both cases.   Table 6. Corrected mirror alignment analysis. (Case 19b Thick Section/Steel Hub)  

In consideration of these results, a  lesson learned is that for a mechanism using thin section bearings with  a combination of tight accuracy, significant loads and temperature variability, design constraints should be  loosened to allow mounting of the bear ing in a material with matching CTE . Moreover,  when the bearing is  designed and analyzed as a rigidly preloaded system, it should truly be rigidly clamped if you are to truly  trust the analysis . Otherwise, analysis results may overlook effects which are difficult or impossible to model  accurately. The modification of the bearing configuration between the Brassboard and the EM/FM1/FM2 is  not believed to have been  detrimental to the design. In fac t, it was expected to be beneficial because the  change reduced predicted bearing stresses.  Had the B rassboard experienced the vibration environment, it  too may have experienced similar behaviors.   Root Cause Investigation   The process of the bearing anomaly investigation was informed by a fishbone diagram ( Figure 12), which  lays out all reasonable causes for the drag torque spikes in the SCS bearing assembly . The diagram is  color -coded to indicate the status of each possible cause. Green indicates that the cause has been  exonerated, yellow indicates that the cause cannot be exonerated until disassembly and inspection of the  bearing assembly, and red indicates that the cause is implicated in causing the anomaly .   The top branch of the fishbone diagram concerns the bearings themselves . Manufacturing defects, poor  design, inadequate analysis, or over -testing could all cause the anomalous torque spikes . The hardness of 150170190210230250270290 -50 -30 -10 10 30 50 70Max Mean Bearing Contact Stress (ksi) Shaft Temperature (C)Effect of Temperature on Max. Mean Hertzian Contact Stress in Bearing CDR Thick Section/Al Hub Thick Section/Al Hub 20 deg DeltaT Thick Section/Steel Hub Thick Section/Steel Hub 20 deg DeltaT
326 14, Orbis 3.0, and MESYS Version 07/2019) at all as -built and tested conditions, and agreed on the  maximum mean Hertzian contact stress  at assembly conditions within a reasonable range of approximately  20 ksi  (140 MPa) . None of the tools predicted the anomalous behavior observed, indicating either the  bearings cannot be properly modeled in the software, or their as -built configuration does not match the  analyzed configuration. Therefore, the analys es performed must be called into question.  

Figure 12: Fishbone diagram 

All our nondestructive evaluation methodologies are all uniquely limited in their capacity to detect possible  defects, so bearing defects and lubrication were considered conditionally exonerated until disassembly of  the bearing set could be performed.   The bearing set design was conditional ly implicated as a cause for the torque spikes. This  set design can  be considered unique or unconventional due to the materials and fits used. Such uniqueness emerged from  the complicated dimensional, environmental, and structural requirements imposed on t he SCS. Until the  bearing set can be properly simulated with the available analytical tools or inspected during a tear -down  procedure, its design must be identified as a leading cause of the anomalous behavior.   Poor assembly technique of the bearings could contribute to the type of anomalous behavior seen in this  bearing set. However, all three SCS units exhibited the anomalous torque spike behavior at exactly the  same position (where mirror was constrained during vibration testing and temperature ramps in TVAC).  Although this paper focuses on the EM bearing set, both the Flight Model 1 and Flight Model 2 exhibited a  similar behavior. The units only differ in magnitude of the torque spikes, where the EM had by far the highest  measured torque spikes. The primary difference in testing between EM and FM1 is how the thermal  environment was controlled in TVAC. Thermal gradients across the entire SCS structure were more closely  controlled for the FM assemblies.   Contamination of the bearing sets was conditionally ex onerated as a cause for the anomalous behavior.  Typically, any damage to a bearing set as a result of Foreign Object Debris (FOD) would manifest as a  behavior isolated to one build of a bearing set. In this case, however, all three bearing sets that were built  and tested exhibited very similar behavior. This cause is conditionally exonerated until a careful  disassembly of one of the bearing sets could reveal some type of FOD generated by the assembly process.