ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
273   

Figure 5 - Left: average temperature & torque vs. speed (2- mm copper disc, 35- mm air gap) ;  Right: measured temperature distribution of spinning dis k (same configuration  as left, at 4500 rpm) 

Temperature measurements have been made using infrared thermometers pointed at a high emissivity  face of the spinning dis k (ensured by coating it with matte black paint) , and torque measurements have  been conducted in parallel.  

It is evident in Figure 5 (left side) that the rises in torque and temperature fall together, as predicted.   Regarding the evolution of conductivity of the dis k as function of temperature , the largest measured  increase above room temperature has been approximately +35 K  (= 55°C -  20°C).  

Figure 5 (right  side) shows a thermal image of the ECTB running in this configuration of maximum heating.  The annular shape of the temperature distribution corresponds to where most of the heating occurs, namely  where the magnetic field has been strongest (directly between the magnet’s pole pieces  that can be seen  left side of the image).  

Assuming the temperature coefficient of resistance to be 0. 00404/K  for copper , this implies a decrease in  conductivity of the spinning copper dis k of more than 14% compared to its value at am bient temperature.  The resulting Lorentz force (and so the torque)  also drop by a corresponding amount. T hus, thermal effects  are not negligible in this context . An adjusted conductivity has been used for the respective simulat ion runs  in the Maxwell and COMSOL , shown in Figure 4 as the “86% sigma” curves (the original “100% sigma”  curves are for a conductivity that is not adjusted). As can be seen, this has resulted in even closer  correlation of the results with measurements.  

Eddy Current Effects  in Reaction Wheels  

A relevant case where eddy currents are an unwanted side effect can be found in reaction wheels  for  spacecraft attitude control . They typically consist of an electric motor driving a metallic flywheel with speeds  up to 6000 rpm and more. Under the presence of a relatively strong external magnetic field, possibly  originating from magnet ic torque rods used as secondary actuators , an additional loss torque due to eddy  currents  in the rotating flywheel may  occur and result in degraded performance of the reaction wheel  concerned.   In order t o quantify this effect, a dedicated test was devised using the Magnetic  Coil Facility at ESA/ESTEC,  which features Helmholtz coil s able to generate magnetic fields up to 7.5 mT. A reaction wheel was  placed  in the cent er of the Helmholtz coils and operated under varying magnetic flux densities. The reaction wheel  assembly  used comprises  a spoked flywheel made from stain less steel. Its maximum motor torque is about  235 mNm,  over a speed range up to 2700 rpm. The l oss torque of the reaction wheel assembly  is typically  in the range of 10 to 1 5 mNm (excluding the effect  of any ambient/external  magnetic fields ). 

 0.000.040.080.120.160.200.24 0.010.020.030.040.050.060.0 0 1000 2000 3000 4000 5000 Torque [Nm]Temperature [ °C] Speed [rpm]Average temp. Loss torque Air gap position  Spinning disk
274      

Figure 6 - Left: Test configuration  (reaction wheel spin axis orthogonal  to the magnetic field vector )   Right: Loss torque vs.  ambient magnetic flux densit y 

For measuring the influence of ambient magnetic fields on wheel performance, the reaction wheel ’s loss  torque  was measured while generating magnetic fields with varying magnitude (following a specific  waveform to ensure that the observed losses are evidently due to the external magnetic field). The test was  repeated for magnetic flux densities ranging from 1.0 to 5.0 mT as well as with different reaction wheel  orientations, i.e. , spin axis parallel and orthogonal to the magnetic field vector.  No measurable effect was  observed when the field direction was parallel to the wheel spin axis. When the field was applied in a  direction orthogonal to the spin axis, a significant  change of th e loss torque was identified , which closely  followed the waveform of the external field vs. time . 

Figure 6 (right  side) shows the dependency of the measured loss torque vs. the magnitude of the ambient  magnetic field. It can be seen that eddy current related loss torques can rise to levels of >100 mNm, i.e.  nearly  half the available motor torque. This happened, however , only for a very significant magnetic field of  5 mT , a level which is unlikely to occur in a real flight situation.  

Moreover, it can be noted that there has been a cubic relationship  between loss torque and flux density .  This has been a surprising result  since, according to Eq.  1, a square relation ship would be expected. It is  assumed that the geometry of the flywheel as well as the spatial distribution and direction of the external  magnetic field play a critical role in the generated eddy current  effects, wh ich is subject to confirmation by  ongoing research.  

Conclusions  

Eddy current related loss torque  effects have been  studied in depth,  both in terms of simulation and  hardware test result s. A nonlinear dependency of the loss  torque on relative motion speed has been  confirmed, which can be  coherently explained by  the combined effect of a counteracting magnetic field   (generated by the eddy currents ) and a decrease of electrical conductivity due to the rise in  temperature by  Joule heating.    The results obtained and the consistency between simulation  output and measurements are promising and  give confidence in the fidelity of the analysis  tools, particularly when mechanical, thermal and  electro magnetic aspects are combined. Ongoing ESA research on the various los s torque components  in  reaction wheel  assemblies  and other space mechanisms will definitely benefit from the  investigation. It will  allow a more accurate predict ion of mechanism performance (especially at higher  speeds ), for example in  the frame of long- term health monitoring as outlined in [1] . 

   020406080100120140 1 2 3 4 5Loss torque [mNm] Ambient magnetic field [mT]2700 rpm   y = 1.4x^(2.9) 1800 rpm   y = =1.2x^(2.8) 1300 rpm   y = 0.9x^(3.0) 800 rpm     y = 0.8x^(2.9)Reaction Wheel  Assembly   Helmho ltz Coils
275 Acknowledgments  

The authors gratefully acknowledge the cooperation and support by many technical groups at ESA/ESTEC,  in particular by the colleagues of the AOCS & Pointing Systems Section, the EMC & Harness Section, the  Materials & Processes Section, the Metrology Laboratory and the Mechanical Workshop.  

References  

1. Bojiloff, D., Häfner, T. & Seiler, R.:  “Health Monitoring for Spacecraft Reaction Wheels”,   Proc. of the 17th European Space Mechanisms and Tribology Symposium, Hatfield, Sept. 2017.  

2. Smythe, W. R.: "On Eddy Currents in a Rotating Disk" ,   AIEE Electrical Engineering Transactions , Volume: 61, Issue: 9, September  1942 .  

3. Gay, Sebastien E. : “Contactless Magnetic Brake for Automotive Applications”,   PhD Thesis , May 2005.
278 R4 Bearing Life -Testing  

The initial test plan envisioned testing 40 brushless DC motors in thermal vacuum bell jar chambers that  were available at NASA M arshall Space Flight Center (MSFC)  [5]. The large number of motors selected  was an attempt to obtain statistically relevant life- test data while discerning  any effects on bearing life  attributed to r ace curvature, speed, temperature, and mode of operation . Additionally, we were interested  in evaluating the differences in performance between all steel and hybrid (ceramic balls)  bearings.  

Motor Selection   Motor selection was driven by the size of the bell jars and the chiller plates that they  could accommodate.   Our focus was placed upon procuring vacuum compatible test motors that could easily incorporate our  ABEC -7, angular contact , R4 size , 440C bearings  and accommodate  the existing test rigs  at NASA MSFC .  Special vacuum and cleanroom proces ses were imposed on the motor structural parts and windings, as  well as, the capability to set  the same bearing preload regardless of race curvature or bal l material (steel  or ceramic). We approached a high- volume production motor supplier willing to accommodate our custom  requireme nts and selected 3- phase brushless DC motors with encoders to provide accurate speed control.  Shaft and housing were modified for our R4 test bearings to establish proper fits over a test temperature  range from ambient to 60 °C. The bearings were preloaded with a wave washer that could be adjusted for  each motor  to meet the target bearing preload of 2.9 ±  0.23 kg ( 6.5 ± 0.50 lb) using a force gau ge. The  2.9-kg preload approximates the mean contact stress of approximately 8.55e 8 Pa (124 ksi) of  flight RWA  steel bearings  (Figure 1) . The motors had strip heaters with thermocouples mounted on t he rear portion of  the housing.  

Figure 1. Test motor showing component parts (strip h eaters and bearings not shown) . 

Lubrication   There was an expectation that the test motors  would have extremely long lifetimes if the bearings  were  nominally lubricated [6]. In an attempt to minimize test time, the bearings  were lubricated with a minimal  amount of Nye Synthetic Oil 2001 . The procedure use d to effect this outcome involved immersi ng an oil  lubricated bearing in an oil-solvent mixture followed by evaporation of the solvent until a very small  meniscus was observed as determined visually under a microscope (Figure 2) . This procedure was referred   to as 7% lubricant slosh. The final bearing weights were measured and the variation found to be acceptable  and typical of bearing to bearing variation in the field.  One exception to performing this procedure was  motor 40 which received a 3%  lubricant  slosh wher e a meniscus was not observed. This exception was to
279 reproduce the lubrication condition present in a prototype test that was run previously in an all steel  configuration.  

Figure 2. Oil meniscus  present in R4 Bearing after 7% lubricant slosh ( Nye Synthetic Oil 2001/heptane) . 

Figure 3. Test motors mounted on chiller plate and installed in bell jar.  

Phase I Testing   The motor s were divided into three sets, mounted on chiller plates, and each set placed in a vacuum bell  jar (Figure 3). Two bell jars labeled Bell Jars #1 and #2 contained 16 motors each and had all steel bearings ,  while a third bell jar labeled Bell Jar  #3 contained 8 motors and had hybrid bearings. Each bell jar contained  motors possessing both 52% and 54% bearing curvatures.  Phase I testing consist ed of all motors running
282 Table 1. Summary of Motor and Bearing Results   Steel Bearings   Motor  Curvature  (%) Failure  (revolutions)  Speed   (rad/s)  Mode  Recovery   22 52 4.36e6  314 (3000 rpm)  biased  No  8  52 5.79e6  105 (1000 rpm)  biased  Full  7  52 9.95e6  105 (1000 rpm)  biased  No  29  52 1.09e9/1.34e9  105 (1000 rpm)  zero crossings  Yes/No  24 54 1.36e9  52.4 (500 rpm)  zero crossings  No 

Hybrid Bearings   Motor  Curvature  (%) Failure  (revolutions)  Speed   (rad/s)  Mode  Recovery   40 54 3.25e9 314 (3000 rpm)  Biased  Suspended   36 52 5.30e9  314 (3000 rpm)  biased  No  33  52 5.37e9  314 (3000 rpm)  biased  No 

Hybrid Motors   Hybrid motors 33 and 36 were shut down to evaluate recovery techniques . The torque signature and  acoustic data for motor 33 were consistent with bearing dr y cage instability as shown in F igures 6 and 7,  respectively. One motor was rested for 37 days then ru n another 45 h ours bef ore suff ering a probable cage  failure. The other motor with hybrid bearings was also rested for 37 days and then heat soake d at 60 °C for  another 30 days when i t failed from probable cage fracture after 7 hours of additional operation (Table 1) .  Neither motor with hybrid bearings has been torn down for inspection to validate the failure mode so as not  to disturb th e remaining motors under test.  

Figure 6. Motor 33 torque signature for dry cage instability.
283   Figure 7. 12-20 dB spikes  between 1400- 2400 Hz associated with dry cage instability on motor 33.  

As mentioned earlier , and as shown in Table 1, motor 40 with hybrid bearings was suspended at 3.25e9  revolutions. This motor was run at 314 rad/s and 50°C  to approximate conditions that were run previously  in prototype testing where the motor in an all steel configuration failed at ~250 million revolutions . Note that  motor 40 was lubricated with a 3% slosh and did not have a visible meniscus. As can be seen in Figure 8,  the Coulombic torque increased as the viscous torques decreased. These coincident events were clear  indications that the lubricant was  depleted and the bearing was  running in a “dry” condition. Further  operation likely would have triggered a dry cage instab ility and eventual cage fracture, and it was for these  reasons that the test was suspended in order to preserve the motor for eventual teardown and inspection.   

Figure 8. Plot of viscous and Coulombic torque versus distance.
284 Steel Motors   The same recovery techniques were applied to motors 7 and 29 with all steel bearing configuration and  which failed at approximately 1 billion rev olution s. They were rested 37 days at ambient and then heat   soaked. Motor 7 reversed direction 5 times and was rested for a total of 227 days w ith no signs of  improvement.  Motor 29 was rested for 37 days then heat soaked unpowered at 60 °C for an additional  30 days with the motor  recovering for approximately 260, 000 revs before failing again. Remarkably, motor  8 with all ste el bearings showed high torques at 579 M revs and was rested at ambient for 37 days. It then  recover ed fully, and is currently healthy at 2.2 B revs while performing zero speed crossings  (Figure 9). 

Figure 9. Typical day before failure (blue), f ailure event (orange) , restart (gold ) and recovery  (purple) . 

Motors in Recovery Protocol   Motor 15 exhibited anomalous torque spikes at 2.23 billion revolutions, was rested for 42 days, and  recovered for 30 days upon which time torque spikes were observed again and the motor rested and  subjected to a heat soak.  Motor 16 also exhibited torque spikes at 2.28 billion revolutions at which time it  was shut down and rested. Motors 15 and 16 were not considered fail ures as recovery was attempted.  In  contrast, motor 32 exhibited behavior consistent with dry cage instability and was shut down and subjected  to a 37- day rest period.  Motors 15, 16, and 32 all possessed 54% curvature.  The results of the recovery  protocol  imposed on motors 15, 16, and 32 were not yet available at  the time of writing this paper as shown  in Figure 5.  

Figure 10. Film thickness versus temperature at 314 rad/s and 105 rad/s.
288 second bearing was depleted at 1 0 years , where the viscous drag fell to near zero. After the onset of  depletion sporadic relubrication events occur due to increased bearing disturbances and temperature.   

Perhaps the most interesting observation of Figure 1, however, is the relative stability through most of life.  Models of lubricant degradation and loss typically imply gradual consumption of oil, yet our observations  show a stable quantity at the rolling and sliding interfaces until rapid loss late in life. This trend in Q is  interpreted as being due to our imperceptibility of the total amount of lubricant in the bearing, much of which  is not actively engaged in the critical interfaces. While this bulk oil is gradually consumed or lost, the oil in  the interfaces is maintained at a constant level that is controlled by the combination of mechanical and  surface forces acting on the oil. These forces are assumed to be constant during spacecraft operations, so  the prevailing thought is that the oil volume does not change and is sustained i n balance with oil reserves  resting outside the active tribological interfaces. For example, this may be in the form of grease reserves  that are channeled to the sides of the ball path.  

Figure 1. A graphic representation of the trends in viscous coefficient that are often observed in highspeed spinning devices . 

A simplified model of the geometry is shown in Figure 2. The Hertzian contact between the ball and the  race creates a capillary, which draws oil into a meniscus. At some distance from the contact are reserves  of oil (held in grease thickener), and this oil migrates between these two bodies by creeping in a th in film  across the bearing surfaces (balls, races, retainer). In addition to the capillary forces of the two bodies, the  oil will be subject to mechanical forces (such as centrifugal acceleration) and other surface forces (such as  thermocapillarity due to temperature gradients  from localized heating).  The direction and magnitude of  lubricant creep will depend on the balance of these forces.   

Figure 2. Schematic of an idealized ball/race contact and its relationship with lubricant reserves  0.000.050.100.150.200.25 0 2 4 6 8 10 12 14Absolute Value of Viscous Slope Operational Lifetime (Yrs)Typical Viscous Slope Trend
289 This demonstrat es that the sustenance of lubricant in the critical interfaces depends on lubricant mobility  on bearing component surfaces. Ultimately, the longevity of critical spacecraft components depends on  management of thin films of oil. Understanding the factors controlling these films is essential to prolonging  life and responding to flight anomalies caused by lubricant starvation.  

Thin Film Flow  

The apparatus shown in F igure 3 A is an adaptation of a device that has been described elsewhere to  measure flow of very thin lubricant films [4-10]. In our work, the substrate is a polished steel coupon, with  characteristics (composition, roughness, etc.) that are similar to bearing components. A fter a thin oil film   (h < 5um)  is cast on to a polished  steel surface , a jet of nitrogen gas is then directed to the coupon surface  causing the oil film to thin over time. F igure 3B shows an optical image of the film during thinning by the N 2  gas jet, where the fluids  film thickness ( h) is continuously monitored with an interf erometer down to   h ~10nm.  The rate at which h declines  over time, known as the thinning coefficient ( β, units of nm *s), is  proportional to the oil’s viscosity and enables  us to determine the viscosity of thin fluid films, a fundamental   property that governs the oil’s interfacial -surface mobility.  

Figure 3. A) Thin Film Flow (TFF) apparatus, B) Thin film interference pattern after blow -off with a jet of  nitrogen gas in the TFF . 

Figure 4 shows an example of the film thickness  (h) vs time during flow from the center of the target area.  In this case, the sample was a 2- µm-thick film of a multiply -alkylated cyclopentane fluid. We find a rapid  decrease in film thickness , with a functional form that appears similar to the film height decreasing in inverse  proportionality to the elapsed time of exposure to the nitrogen jet. Figure 5 is the same data, plotted as the  inverse film thickness  (1/h) vs time . This plot provides a better demonstration of the inverse proportionality,  where the slope is th e thinning coefficient, β. We also find that there are two distinct regimes of viscous  behavior. The first, with larger film thicknesses and early time of the experiment, represents the bulk  properties of the fluid. The second regime represents the interfacial properties of the thin film, where  interfacial forces that exist between lubricant molecules and the steel surface influence fluid mobility . In this  case, the transition between these two regimes occurs at a film thickness of ca. 45 nm. After the transition  to the interfacial regime, the viscosity was observed to increase by approximately  32%.
290   Figure 4. Plot of film oil film thickness vs time during blow -off using the TFF  

Figure 5. Plot of the data shown in Figure 4, using reciprocal of film thickness on the y -axis to show the  inverse proportionality with time  

These changes in viscosity and flow rate are important, because the process of resupply to the tribological  contacts may be slowed as lubricant consumption proceeds and the scarcit y of free oil leads to reduced oil  film thicknesses.  

Tribometry and Viscosity of Worn Lubricant Films  

Another significant reduction in oil mobility during operational use may be caused by changes in oil  composition. As the lubricant is worn in a tribological contact due to mechanical stresses and chemical  reactions, some molecules are broken into smaller components while others are polym erized into larger
291 molecular weight varieties. The lower molecular weight products may evaporate more readily, and the larger  products tend to accumulate in the  vicinity of the contacts.  This occurs with potential increases in the  effective viscosity, and may affect both resupply of the contact and elasto- hydrodynamic film thickness.  To  study the effects of lubricant degradation and changing composition on surface flow in the contact region,  we have mechanically worn the lubricant using a ball on disk tribo meter in a vacuum chamber. This  inhouse developed instrument, shown in Figure  6, uses steel coupons similar to those shown above [7, 9,  11]. A multiply -alkylated cyclopentane oil with phosphate additives was used at 24° C and background  pressure of 5 x 10-6 torr. With a load of 2.6 N , the contact stress was approximately 580 MPa  and the  diameter of the Hertzian contact area was 0 .6 mm. The rotational speed was 90 rpm (boundary conditions).  However, in some cases, a lateral oscillation was applied to lift the contact into EHD conditions.  

Figure 6. High vacuum ball -on-disk tribometer  

After running under vacuum for 30 hr , samples were transferred to the TFF apparatus for measurements  of vis cosity.  Three separate tests are shown in the top of Figure 7. The first two (Un88 and Un89) were  performed at low speed, and Un 90 at high speed. In each case the contact path is visible as a distinct line  in the oil film, caused by a defect in the otherw ise gradual topography of the film thickness due to pinning  and scalloping.   

After blowing with nitrogen gas for over 1 hour, the displaced oil film patterns are shown in the bottom of  Figure 7. In the two low speed examples, we find distinct regions of impeded oil flow that range from 3.1 to  5.2 mm across. These tracks are much larger than the diameter of the Hertzian contact, so they are not  caused by changes in the substrate surface roughness.  Instead they are related to the size of the ball/flat  meniscus, where products of the lubricant degradation have been carried from the Hertzian contact and  deposited in the region of oil reflow. In the high- speed case (Un90), bands of impeded oil flow are not easily  detected and is likely because the EHD contact was not sufficiently stressful on the lubricant to create  substantial wear and lubricant degradation polymerization  products.