ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
88  

Figure 14 . Scan Drive Motor Current 1 Year History  

Conclusions  

From the years of effort related to the SNPP ATMS scanner bearing system , the following recommendations  and observations can be offered:   1) Use of Teflon toroids in space bearing applications has to consider  the range of motion of the  mechanism . Based on this experience, uni -directional applications for toroid use cannot be  recommended. Applications with regular directional rev ersals may be  acceptable and even  desirable, but must be based on a good and representative life test program.   2) Uni-directional motion will result in eventual dry running of the toroid/ball interface leading to the  premature failure of the toroid separators .  3) Bearing life test programs must use flight like hardware, not only for the bearings but also for the  drive motors , encoders, electrical drivers  and telemetry sensors . A flight -like Engineering  Development Unit  of the mechanism in other words . This would allow for a one to one  comparison of life test data to on- orbit telemetry . Not doing so, results in a very difficult analysis  task of any on- orbit data when anomalies occur.   4) Motion reversals are very effective in preventing or arresting the wear of bearing toroids . It is  evident from the bearing test units and on- orbit data that reversals eventually lead to acceptable  bearing performance . Depending on the state of toroid wear, initial reversal results may include  ‘spiky’ behavior . However, data shows that e ventually the bearings settle down into a near normal  state . The SNPP on- orbit torque levels are now subsiding to near early orbit values, although with  some spiky behavior .  

Acknowledgements  

Every successful major project results from the combined effor ts of many people. It is with great regret we  don’t have the space to list all of the many people who contributed time, effort, and talent to the ATMS Scan Drive  program. Nonetheless, most of those who participated in the Scan Drive Life Extension program  would agree that the following people were key to the suc cess of this effort . Goddard Space Flight Center:   Mark McClendon, Chris Hoffman, Robert Lambeck , Bruce Guenther , Jason Sturgis, Tom Manson, Harry  Solomon, Tamara Oconnell , Michael Dube, Ed Dobbins , Dan Helfrich ; NOAA: Brennan Nowak,  Jim Waters,  Ninghai Sun;  Northrop Grumman Electronic Systems: Mike Landrum, Paul Schan, Victor Jacobo, Kent  Anderson; Aerospace Corp: Tina Gentry, Neal Baker , Marc Wigdor , Joe Pope; Lincoln Labs: Vince Leslie  

   Initial Reversal   Start of Daily    Reversals
89 References  

1) Advanced Technology Microwave Sounder on NPOESS and NPP, Christina Muth, Paul Lee,  Sergey Krimchansky, James Shiue, Allan Webb. NPOES IPO Information center, No. 06125, 06-16-2004, http://140.90.86.6/IPOarchive/SCI/sensors/ATMSDraftBriefv4_AlanWebb.pdf

2) Design of the ATMS Scan Drive Mechanism, Curtis Allmon , Dave Putnam; Proceedings of the 38th  Aeros pace Mechanisms Symposium, Langley Research Center, May 17- 19, 2006   3) Zhao- Zhu Zhang, et.al, “Effects of Various Kinds of Fillers on the Tribological Behavior of  Polytetrafluoroethylene Composites under Dry and Oil -lubricated Conditions”, J. of Applied  Polym er Science, V80, #11, Jun 2001, pp. 1891- 1897.
93 side bearing failure investigation are presented along with the development of the likely ro ot cause and  failure recovery.  

Figure 1 .  Representative ISS Distillation Assembly cross- section taken from reference 8  

Bearing Design and Operating Conditions  

Table II lists  the nominal operating conditions and the bearing characteristics for the 50- mm bearings.  The load on both bearings comes primarily  from the preload wave spring. The outer races use an  interference press fit into the drum end walls and the inner races slide over the shaft with a small  clearance.   Table II .  Representative Application Bearing Parameters and Operating Conditions   Parameter  Value or Condition   Outer Diameter, OD  ~80 mm   Inner Diameter, ID  ~50 mm   Width, W  ~16 mm   Ball Size, D  ~9 mm   Ball Material  Si3N4  Race Material  60NiTi   Cage  Snap fit, polymer   Lubricant  Lithium based grease   Ball-Race Stress Limit  ~3.1 GPa   Ball-Race Mean Stress  ~1 GPa   Axial Preload  ~200 N   Radial Load (terrestrial)  ~100 N/bearing   Speed  100-300 rpm   Environment  Warm, highly acidic aqueous solution   Ambient Pressure  Slight vacuum
94 The operating loads and speeds are low  for a 50- mm bearing. In this application, the bearing size was  largely dictated by the need to accommodate a large, stationary hollow shaft that houses electrical and  fluid flow passageways into and out of the drum assembly. Other than potentially severe vibration loads  (~3 g or higher) during a rocket launch, the operating conditions are characterized as mechanically  benign and chemically aggressive. Since the 60NiTi alloy had previously been shown to be impervious to  the acidic wastewater, the premature bearing failure was unexpected.  

Hardware Observations and Other Forensic Evidence    The presence of nickel -titanium wear debris trapped by the DA filter screens indicated that bearing wear  had occurred prior to failure. Following test termination, the hub- side end wall plate of the DA was  removed making one face of the bearing visible. Figure 2 shows the view from the hub- side of the DA  looking towards the motor -side.    Based upon this image alone, one plausible preliminary failure scenario is that the inner  race fractured  and caused more generalized damage and wear as the test continued. A close up image of the inner race fracture, shown in Figure 3, offers additional clues to the behavior of the 60NiTi.  

Figure 2 .  Hub-side view of failed ISS Distillatio n Assembly centrifuge   bearing . Note cracked inner race and ejected wear debris  

The inner race appears to have sprung radially outward from the DA shaft to the left side of the crack and a significant gap has opened between the faces of the crack. This behavior suggests that the inner ring  had been under considerable residual stresses and these may have been a factor in the bearing failure.    Indeed, the bearing manufacturing process for 60NiTi includes a rapid water quench step following heat  treatment at 1000 °C [9]. This rapid cooling results in the rise of residual thermal stresses. Unlike steels  and other metals in which such residual stresses can be relieved by soaking the parts at elevated  temperatures, NiTi has no mechanism for internal atomic leve l slip. The residual stresses that remain  after the quenching step cannot be relieved through additional thermal treatments and these stresses may be a factor in undesirable effects like ring cracks [10, 11].
95  

Figure 3 .  Close -up view of cracked inner  race 

A similar bearing ring crack was observed during prototype bearing development. A bearing had been  successfully tested, removed from the test rig and inspected. The bearing was in as -new condition and  was placed in storage. Several weeks later, the inner ring developed a crack similar to the one observed  in the present case. Figure 4 shows that inner ring crack.  

Figure 4.  Photo of an early prototype bearing that developed an   inner race crack during post -test storage    In this instance, forensic a nalyses showed that the early prototype 60NiTi material used in this bearing  contained significant non- metallic inclusions and other internal flaws that could have triggered the crack.  In addition, the bearing was damaged during final machining of the insi de surface of the inner ring.  Adjacent to the fracture location, a low spot and a deep gouge are observed. This damage was a
98  

Figure 5.  End views of DA drum shown during bearing installation. Hub side   bearing housing structure (a) differs significantly from motor side (b).  

Figure 5(c).  Close up end views of hub -side of DA drum showing   how the spoke design allows ready access for assembly.   Photo taken from [7].  

For the hub- side bearing the situation is more nuanced. The known interference fit is larger, 0.0011 in    (28 µm). This level of press fit into a circumferentially symmetric and relatively rigid housing would be  expected to reduce the bearing diameter by about 0.0009 in (23 µm). This level is greater than the known  bearing internal clearance (0.0008 in (20 µm)) resulting in the bearing having no internal clearance and  indeed the balls and the inner race would be under a preload by the outer race. This in itself is an  undesirable condition that may factor into the hub- side bearing failure. However, the circumferentially  non-uniform spoke structure supporting the hub- side bearing housing may cause additional problems in  that the installed bearing outer race may not be compressed uniformly.  

To estimate the magnitude of the deformations of the hub- side bearing outer race, a finite element model  was developed. Figure 6 displays the model result for the axial deformation of the hub-side three- spoke  structure under a greater than normal 200- lb (890- N) axial load. The normal axial load for the DA is only   40 lb (178 N) . The 200-lb (890- N) case was run to determine whether axial deformation arising from the  spring preload could possi bly distort the hub- housing geometry. The result was clear. Even under  exacerbated axial load the housing shape remains round. It was noted, however, that significant axial deformation of the spokes occurs even under the 40-lb (178- N) design load. This def ormation effectively  reduces the axial preload through reduction of the distance between the outer races of the hub and
99 motor -side bearings. Thus if the wave spring height reduction is used to set preload, it is likely that the  installed preload will be lo wer than expected by about 10%.  

Figure 6 .  Finite element deformation of the DA hub side structure in response   to a large ( 200-lb (890- N)) axial load. Note that the hub bore remains round  

Figure 7 .  Finite element deformation of the DA hub side structure into which a   bearing is installed with a heavy (0.0011 in/28 µm ) press fit. Note that the three   spoke locations provide high rigidity leading to an irregular bore shape
100 The second case examined was for a heavy interference press fit (0.0011 in (28 µm) diameter difference)  of a round bearing into the circular hub- side housing bore. Figure 7 shows the result.  

At the spoke locations, the bearing outer race diameter is reduced by approximately the same magnitude  as the bearing clearance and the less  rigid housing sections between the spokes allows the bearing  housing to take on a tri -lobe shape with radial variation of approximately 0.0002 in (5 µm). Such  roundness variation combined with the loss of internal clearance could be responsible for the ra pid  bearing failure. To corroborate such a root cause theory, however, requires experimental verification.  

Bearing Testing -Experimental Corroboration of Failure Theory  

The failure of the hub- side bearing in the DA centrifuge after only 200 hr of operation was not anticipated.  The operating speed was low (200 rpm) as were the applied external loads. Short -term testing of 60NiTi  centrifuge bearings had not surfaced any problems and long- term sub- component 3 ball -on-rod rolling  contact fatigue tests suggest infinite life at the design stress levels [13]. However, no long -term life test of  a full -scale DA centrifuge bearing had been conducted prior to the failure. Therefore, any bearing tests  carried out to corroborate a root cause theory must achieve two goal s. First, a bearing test under similar  conditions as the failure mode must produce a similar failure in a similar test period. Second, the full - scale, long- term bearing life test under normal design conditions must be done to show that if the factors  that contributed to the early failure are ameliorated, long bearing life ensues.    To achieve these two test goals, a series of experiments were conducted in the NASA Glenn DA bearing test rig. This test rig, described in an earlier report on the bearing design and fabrication process, mimics the centrifuge speed, load and other mechanical conditions and can be run continuously and unattended  [9]. It does not include the aggressive chemical (wastewater) environment but given 60NiTi’s excellent  corrosion resistanc e, the absence of the warm wastewater is not believed to be a failure contributor. The  rig is shown in Figure 8.  

Figure 8 .  NASA Glenn’s DA centrifuge bearing test rig. The present failure corroboration   tests are run without water and heat.
101  

Figure 9 .  Shims used to deform the bearing outer race to match   operating conditions in the DA that lead to the premature failure  

The first test undertaken was to duplicate the hub- side bearing fit conditions that are believed to have  caused the premature bearing failure. This could be achieved by manufacturing a new test rig bearing  housing with the same three spoke structure and fit as the DA. Instead, a simplified approach was used that did not require extensive fabrication. The existing cylindrical housing was  retained and three small  pieces of stainless steel shim stock were placed between the bearing outer race and the housing. These three shims were approximately the same circumferential length and depth as the DA spokes. The shim thickness was selected to deform the bearing outer race approximately the same magnitude as the DA  hub side housing. The use of shims is not an exact simulation of the physical race deformations in the DA  but it is a fair approximation of the clearance reduction and tri -lobe shape v ariation. Figure 9 shows a  photograph of the bearing outer race and the three shims used for the deformation experiment.    Prior to bearing installation, the shims were affixed to the bearing outside diameter at three equally space  locations (120 ° apart). T he bearing housing was warmed to allow a slip fit of the shimmed bearing so as  to ensure that the shims were not dislodged. The second test bearing used (simulating the motor -side  bearing) was fit normally into the housing with no shims. Despite the shims,  the bearings turned smoothly  and quietly with low torque. The test began and was monitored for torque once daily.  

After 150 hr of test time the bearings continued to turn smoothly. At 157 hr an audible noise was noted  and the torque sensor indicated a measurable torque ripple. The test was stopped and the bearings were removed, disassembled and inspected. The shimmed bearing had suffered from fatigue damage. The grease was discolored and wear debris was observed. The fatigue damage resulted in deep racewa y pits  and spall craters. These were located in the regions that overlapped the shims. The other test bearing that was not shimmed showed no signs of damage. Figure 10 is a close up photograph of the raceway  surface of the shimmed outer race after test. Si gnificant surface pitting and spall regions are present. The  inner race also showed signs of surface distress (small pits) but it was uniformly distributed around the raceway and less severe.  

In the largest spall crater, approximately 4 mm in diameter, cr oss-race cracks had developed. In some  places these cracks had propagated to the outer surface of the race as shown in Figure 11.
102  

Figure 10 .  Raceway pit developed during shimmed   bearing test after 157 hr of run -time 

Figure 11 .  Crack originating in the fatigue spall on raceway has propagated to the   outer surface of the race after 157 hr of run -time  

If this test been allowed to run longer it is reasonable to expect the damage to progress possibly until the race cracked through completely. Upon reflection, it is possible that the failure path for the DA bearing was surface fatigue followed by pit growth, crack formation and propagation and finally, race fracture. At the onset of the investigation, one theory was that the raceway residual stresses had led to ring fracture  and subsequent bearing wear and debris generation. Based upon the shimmed bearing experiment, it appears that ring fracture may have been the final st ep, not the initiation step.  

The second test undertaken was to prove that the 60NiTi bearings have the ability to run for a long- term  test without failure, providing they are installed with fits and clearances well accepted as good industry  practices. To conduct this test, a pair of flight quality 60NiTi bearings was carefully inspected, cleaned and lubricated. The pair was then installed into the NASA Glenn test rig with fits within the specified design range. The inner races had a clearance (shaft to bearing diameter) of 0.0005 in (13 µm) and the  outer races had a slight interference fit of 0.0003 in (7.6 µm). In our test rig, the outer race housing does  not have spokes, thus any outer race deformation is circumferentially symmetric.  

•Crack formed at deepest part of  pit, propagated to outside  surface of race.
103 In addition, an accelerometer was installed on the stationary inner shaft to monitor the rig vibration. Once  each hour the full frequency range (sub- hertz to 100 Hz) is collected and stored. Comparison of these  spectra over the course of the test allow for careful health moni toring. Any vibration growth at relevant  frequencies (e.g., at the run speed, ball pass frequencies, cage frequency) would be cause for bearing  inspection. Figure 12 shows a representative cascade plot of hourly vibration spectra collected over a  three- day test period.    The cascade vibration plot provides a convenient visual method to judge bearing condition. Each peak in the spectra corresponds to different aspects of the test rig and test bearing. For example, vibration at the rig speed (3.2 Hz or 200 rpm ) would arise from any anomaly associated with the drive motor, couplings or  drive shafts. Vibrations generated by the test rig support bearings arise at 22 Hz owing to the number of rolling elements and internal geometry of these bearings. The test rig st ructure has some activity at   37 Hz. A key test bearing frequency for health monitoring is 24 Hz. This frequency represents the ball pass frequency on the inner race. Slight roundness and other variations will lead to vibration at this frequency. Any signif icant increase in the amplitude at 24 Hz could signal raceway surface damage in the  ball path. Thus the cascade plot serves as a snapshot of bearing health.  

Figure 12 .  Cascade plot of hourly test rig vibration spectra captured during a three -day test  period.  Overall vibrations are very low and consistent  

Prior to the start of this second test, the bearings were disassembled, solvent cleaned, visually inspected,  reassem bled and greased (30% fill of the free- volume) with a lithium complex synthetic grease. The  bearings were then installed into the test rig and the pr eload spring was adjusted to 90- lb (400 -N) preload  to match the preload level generally recommended by bear ing manufacturers for this size bearing [12].  This preload is higher than the application design preload (40 lb / 178 N ) but that lower load level was  selected based upon the more modest static load capability of the prior bearing material made from a softer cobalt alloy [9]. Further, the higher load level may offer an acceleration of any fatigue failure  modes. Thus, if the bearings indeed survive the life test, the root cause theory and recovery path is better substantiated.    The life test was begun and vi bration data collected. At first there was a noticeable torque ripple  measured. After a few hours of run time the torque ripple subsided and the cascade plot stabilized
104 probably because the grease had become more uniformly distributed in the bearing. The l ife-test period is  5000 hr. Once during test period (1075 hr) a bearing inspection was conducted. Figure 13 shows images  of one of the bearings during the interim inspection.  

During the inspection the races were examined under 10x optical magnification and appeared to be in  excellent condition. They exhibited no surface pits, scratches or wear tracks. Also, the retainer was unworn and the grease color and consistency was the sam e as the beginning- of-test. Following the  inspection, the bearings were solvent  cleaned, re- greased and installed back into the test rig. After a few  hours of operation the torque ripple subsided leaving a vibration spectra that resembled one from before the inspection.  

Figure 13 .  Photograph of the test bearing during an interim  inspection at 1075 hr. Overall  condition of the bearing races, balls and retainer was unchanged from beginning   of test. The grease consistency and color was also unchanged  

The life -test concluded at 5000 hr because it met the project requirement. More im portant, the successful  life test helps to corroborate the root cause theory by showing that a set of bearings, when properly fit into the machine structure can operate without failure for long periods.  

Root Cause Failure Theory and Recovery Plan    The forensic examination of the failed bearing, subsequent measurements of the related machine hardware, and corroborative bearing tests supports the root cause theory that an overly tight interference  press fit and a non- symmetric housing structure led to surface fatigue damage to the 60NiTi bearing. The
105 surface damage then progressed until the inner race cracked through its thickness and the operation was  suspended. To avoid this failure mode the following remediation steps are indicated.    Selective fitti ng of the bearings, especially in the hub- side DA housing, should be done to seek a fit in  which the bearing outside diameter and the hub bore diameter are as closely matched as possible. The design calls for a range of 0.0001 in (2.5 µm) clearance to 0.0006 in (15 µm) interference. Another  approach is to manufacture bearings with varying diameters to accommodate future DA machines with less or no interference thus avoiding the deleterious effects of outer race distortion and loss of internal clearance. Fur thermore, bearings with larger internal clearances can be used so that any outer race press  fit will not result in a significant or even complete loss of internal ball -to-race clearance. Conversely,  bearing designs less sensitive to internal clearance vari ations, such as angular contact bearings, could be  considered.  

Concluding Remarks    NiTi bearing technology is still relatively immature. Therefore, when problems such as this unexpected  bearing failure arise, it is natural to infer that the new material i s a cause. Indeed in this case, the low  elastic modulus of 60NiTi compared to steel results in more pronounced race deformation due to hub bore interface effects like shape and diameter interference. It is possible that a similarly dimensioned steel bearing would not have failed because the outer race deformations would have been smaller.    For these reasons, it is essential that extra care be taken when using 60NiTi bearings to ensure that all industry accepted engineering and appli cation practices are foll owed. Going forward, it is likely that the  experience base for the use of these materials will grow and this will support the emergence of new best  practices. By adopting these practices, the mechanisms community will be able to capitalize on NiTi’s  excell ent corrosion resistance and static load capability properties.  

Acknowledgments    The authors wish to thank Anita Howard, Walt Schneider and Mike Perry of the NASA Marshall Space Flight Center for their thoughtful technical input. We thank Walt Wozniak and Steve Miller at the NASA  Glenn Research Center for their technical efforts to build the test hardware to corr oborate the root cause  theory. Finally, we thank NASA Glenn’s Penni Dalton for her steadfast project management support and  for standing behind our progress despite setbacks  and other challenges. We have benefitted from the  perseverance and dedication of these colleagues.