ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
Figure 1. Schematic of Mini Traction Machine  The sliding speed is defined as  = |−| (1) where u1 and u2 are the speeds of the two surfaces in respect to each other. The entrainment (rolling) speed  or velocity is then  =		  (2)  With the slide to roll ratio (SRR) being the ratio of the sliding speed to the rolling speed of a contact.   =		 	 (3)  When the slide-roll ratio is between 0 and 2, the ball and disc will rotate in the same direction relative to the  contact. At SRR of 0, the contact sees pure rolling while at 2 it is in pure sliding with the ball held stationary  relative to the disc. While in this range of SRRs, any increase in the sliding speed causes a proportional  increase to the entrainment speed U. This increase in the sliding speed also creates a thicker film separating  the surfaces which will lead to less contact and a stronger EHD film [8].  For this research, in order to keep the tribological contact in a mixed film condition, the use of counterrotation will be employed. Counter-rotation drives the ball and disc in opposite directions independently.  Through operation of the MTM in opposite directions relative to the contact, the rolling/sliding speeds can  become large while the entrainment speed stays fixed. This allows for very high sliding speeds with an  entrainment speed at the contact that can theoretically be 0. The advantage to performing tests in this  manner is to eliminate the effect of an EFD film creation and allows for additives and surface chemistry to  be tested in either boundary or mixed film lubrication regimes.  The film thickness for Mixed Film Lubrication was calculated as follows.  ℎ ≈ 1900 η (4)  where η is the Dynamic Viscosity and U is the Entrainment Velocity.  In previous work by the author and when considering other scuffing tests, it was noticed that most  simulations use a load stage progression. This can be seen in the FZG, 4-Ball EP, SRV, OK Load Test,  and Timken tests. The testing in this work will be done using progressive speed as opposed to a load stage  test. Part of the reason for this is because the higher sliding speeds will allow for a more aggressive wear  rate to help differentiate the efficacy of the additives. Another reason that may be more important is the fact  NASA/CP—2018-219887 72
Figure 2. Mixed Film Lubrication Wear Rate at 20°C  In atmospheric conditions, the wear rates for 52100 steel was consistently lower than the 440C. Under  vacuum conditions, this still holds true for the sample with 0% lead but the samples with 3% and 5% lead  produced lower wear rates with 440C than 52100. Interestingly enough, when looking at the average wear  scar there is no correlation to the disc wear volume of the wear rate. Until recently, most published papers  used the average wear scar to compare efficacy of wear additives and performance in testing. Using this  measurement simply gives you a dimension of the worn area with no indication of how much material was  removed. The results from this study as well as those presented by St. Pierre [9] have illustrated that twodimensional wear measurements cannot be relied on to understand what is going on in a mechanism or  tribological contact. By using 3D profilometry, a deeper understanding of what is going on can be attained.  The results in Tables 4-6 and Figure 3 are for the multiply alkylated cyclopentane with 0%, 3%, and 5%  lead naphthenate when tested in a mixed film regime, at 75°C on the SRV under both atmospheric and  vacuum conditions.  Table 4. Wear Performance for MAC fluid with 0% Lead at 75°C  Table 5. Wear Performance for MAC fluid with 3% Lead at 75°C  00.050.10.150.20.250.3 0% Lead 3% Lead 5% LeadWear Rate (µm3/mm)Mixed Lubrication Wear Rate 20°C  440C Atmosphere 440C Vacuum 52100 Atmosphere 52100 Vacuum Material Temp ( °C)Environment Specimen Avg Wear Scar (mm2)Disc Wear Volume (μm3)Wear Rate (μm3/mm) 0% Lead 75 Atmosphere 440C 1.949 1,080,000 0.50 0% Lead 75 Vacuum 440C 2.286 962,688 0.45 0% Lead 75 Atmosphere 52100 1.978 263,796 0.12 0% Lead 75 Vacuum 52100 2.426 276,480 0.13 Material Temp ( °C)Environment Specimen Avg Wear Scar (mm2)Disc Wear Volume (μm3)Wear Rate (μm3/mm) 3% Lead 75 Atmosphere 440C 2.380 977,730 0.45 3% Lead 75 Vacuum 440C 2.853 691,200 0.32 3% Lead 75 Atmosphere 52100 3.987 622,811 0.29 3% Lead 75 Vacuum 52100 2.216 259,200 0.12 NASA/CP—2018-219887 74
Table 6. Wear Performance for MAC fluid with 5% Lead at 75°C  Figure 3. Mixed Film 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 allows 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 atmospheric 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. As this SRV testing is a pure sliding test, the lower wear resistance for the 440C at 75°C  across all concentrations 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 half the wear at 75°C, approximately the same level of wear at  a 3% loading of lead, and at 5% lead the wear rate was higher for samples tested in atmospheric conditions  while the wear under vacuum was slightly lower.  Material Temp ( °C)Environment Specimen Avg Wear Scar (mm2)Disc Wear Volume (μm3)Wear Rate (μm3/mm) 5% Lead 75 Atmosphere 440C 3.003 756,000 0.35 5% Lead 75 Vacuum 440C 1.874 388,800 0.18 5% Lead 75 Atmosphere 52100 3.335 612,415 0.28 5% Lead 75 Vacuum 52100 1.842 162,000 0.08 00.10.20.30.40.50.6 0% Lead 3% Lead 5% LeadWear Rate (µm3/mm) Axis TitleMixed Lubrication Wear Rate 75°C  440C Atmosphere 440C Vacuum 52100 Atmosphere 52100 Vacuum NASA/CP—2018-219887 75
MTM (Rolling/Sliding Contra-Rotation)  The results in Figures 4-6 are for the MAC cyclopentane with 0%, 3%, and 5% lead naphthenate when  tested in a mixed film regime rolling/sliding configuration on the MTM at 50°C and at 150°C. These plots in  Figures 4 and 5 illustrate the traction coefficient over time during a mixed film scuffing wear test. In Figure  6, the wear rate of each material is shown. This wear rate was determined by the total wear volume (µm3)  per millimeter traveled in the test. In Figure 5, a benchmark sample is added to illustrate a material that was  developed for a non-vacuum application that illustrates exceptional anti-scuffing performance. This is a  target to work towards for anti-wear performance improvement of the MAC lubricants for space  mechanisms.   Figure 4. Mixed Film Progressive Speed Scuffing Test at 50°C  Figure 5. Mixed Film Progressive Speed Scuffing Test at 150°C  0.030.040.050.060.070.080.09 0 200 400 600 800 1000 1200 1400 1600 1800Traction Coefficient Time (sec)Mixed Film Scuffing (52100 Steel, 1.25 GPa, 50⁰C) 0% Lead  3% Lead 00.020.040.060.080.10.120.14 0 500 1000 1500 2000 2500 3000Pennzane Mixed Film Scuffing (52100 Steel, 1.25  GPa, 150⁰C) 0% Lead  3% Lead  5% Lead  Benchmark NASA/CP—2018-219887 76
In general, anti-wear and EP additives are thought to reduce friction in a lubricated contact however in  these scuffing wear tests, the friction for all the samples with lead naphthenate are generally higher in  friction. Although the friction was higher, this did not create larger wear rates. The higher friction in the  contact is caused by the formation of a lead tribofilm on the ball and disc. This tribofilm of lead formed on  the disc surface while protecting the contact also increases the friction against the contact and tribofilm on  the ball. This should not be considered a negative effect but instead illustrates the importance of run-in to  allow lead naphthenate to chemisorb and physiosorb on a metal surface.   Figure 6. Mixed Film Scuffing Wear Rate Comparison  In Figures 7-9, the wear profiles for the 0%, 3%, and 5% lead are shown from one section of the tested  disc. The wear pattern of the experiment using the 0% lead is a typical rough worn groove with indications  of adhesive wear along the contact area. The 3% lead experiment shows a much shallower wear depth and  overall a smoother wear pattern. The 5% lead experiment is the most interesting as there was almost no  wear below the steel surface. In this experiment, it can be seen from the profile analysis that a 2 to 4-µm  layer has been created on the surface of the steel. It is this surface that was then wearing away during the  experiment. Upon further investigation using X-ray Spectroscopy, it was determined that this tribofilm that  had been created on the surface of the steel was primarily composed of lead.  9.67 2.42 0.797.79 1.22 0.36 024681012 0% Lead 3% Lead 5% LeadWear rate ( μm3/mm)Scuffing Wear Rates 50°C 150°C NASA/CP—2018-219887 77
Figure 9. MTM Wear Profile of 5% Lead Naphthenate  From these results, we can see that there is a transition point in the concentration of lead in these MAC  fluids and how it reacts with the surface metal to form an anti-wear tribofilm. It should be noted, that this  phenomenon was not seen in the SRV testing. 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/sliding which will promote  a tribofilm to be created in a fashion much more similar to bearings and gears. This agrees 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. 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  higher contact stresses will create additional wear which will react and consume available lead naphthenate  making it unavailable to create a protective tribofilm. From this research, the following conclusions were  found.   NASA/CP—2018-219887 79
References  [1] Jason Galary. Rolling Wear and Fatigue in Lubricated Contacts. PhD thesis, University of Massachusetts at Dartmouth, 2018. [2] H Blok. Gear wear as related to viscosity of oil. JT Burwell, Jr., American Society of Metals, Ohio, pages 199-277, 1950. [3] HC Mougey and JO Almen. Extreme pressure lubricants. Proc. API, Refining Div, pages 76-94, 1931. [4] FF Musgrave. The development and lubrication of the automotive hypoid gear. J. Inst. Pet, 32:3244, 1946. [5] Jason T Galary. Synthetic gear lubricants go green. Gear Solutions, 8(84):24 31, 2010. [6] Jason Galary. Study of wear properties of environmentally friendly lubricants for gearing applications as a function of film thickness transition. In Environmentally Considerate Lubricants. ASTM International, 2014. [7] LD Wedeven. Traction and film thickness measurements under starved elastohydrodynamic conditions. JOURNAL OF LUBRICATION TECHNOLOGY, TRANS. ASME, Series F, 97:321-329, 1975. [8] Bernard J Hamrock and Duncan Dowson. Ball bearing lubrication: the elastohydrodynamics of elliptical contacts. 1981. [9] St. Pierre, Nicole. “Use of 3D Optical Profilometry to Differentiate Between Additive Chemistry” STLE Annual Meeting, 2017 [10] Didziulis, Stephen V., and Paul D. Fleischauer. "Chemistry of the extreme-pressure lubricant additive lead naphthenate on steel surfaces." Langmuir 7.12 (1991): 2981-2990. [11] Carre, D. J., P. A. Bertrand, and J. R. Lince. Lead naphthenate additive tribochemistry in hydrocarbon oils . No. TR-2002 (8565)-2. AEROSPACE CORP EL SEGUNDO CA LAB OPERATIONS, 2001. [12] Peterangelo, Stephen C., et al. "Improved additives for multiply alkylated cyclopentane ‐based lubricants." Lubrication Science 25.1 (2008): 31-41. [13] Marc Ingram, Clive Hamer, and Hugh Spikes. A new scuffing test using contra-rotation. Wear, 328:229 240, 2015. NASA/CP—2018-219887 81
Motivation and Background  Ball bearing systems in space mechanisms are usually composed of two angular contact or deep-groove  bearings, which are preloaded along their axial direction in order to ensure a sufficiently high stiffness, the  required precision of rotation, and a stable friction torque (within acceptable limits). The bearing preload  can be applied using two different basic methods: hard or soft preload. In the hard preload method, a  deflection is forced on the bearing pair by rigid mounting parts, corresponding to the desired preload  magnitude (knowing the load-deflection characteristic of the bearings). The soft preload method relies on a  compliant element, typically some spring-type component, to apply an axial force on the bearing pair  corresponding to the controlled deflection of the compliant element (i.e., knowing its load-deflection curve).  Both methods have their pros and cons [3] that will not be further detailed here for brevity. In fact, both  methods are used in space mechanisms, and they imply different approaches for modelling and simulation.  A schematic overview of both methods is given in Figure 1.  Figure 1. Hard and soft preload methods in ball bearing systems [3]  When an axial load is applied on a bearing pair (during launch or normal operation), it is generally shared  between the two bearings, increasing the total load on one of them and gradually off-loading the other  bearing. When a bearing gets completely off-loaded, the balls lose the controlled contact with the races, a  phenomenon commonly known as gapping. This effect does not happen in the same way for the two  different preload methods.   Hard preloaded bearings are pressed together with a displacement   corresponding to the preload  magnitude on the load-deflection curve of an individual bearing. For reaching the onset of gapping in a  bearing, it must be displaced by the same amount in the opposite direction, causing an axial load  corresponding to 2	∙  in the other bearing. Assuming a relationship between axial bearing load and  deflection according to Equation 1 (see e.g. [4]) and knowing that the entire load will then be carried by the  remaining bearing, the off-loading or gapping force may be estimated with:  where  is a bearing specific stiffness coefficient. The static behavior of a hard-preloaded bearing pair can  be visualized according to Figure 2. When gapping has occurred in either direction, the load-deflection  curve follows that of an individual bearing. Both characteristics sum up in the operating range without  gapping. Therefore, the linearized stiffness around the preload point (origin in Figure 2) can be  approximated by twice the stiffness of a single bearing at the preload magnitude.   =∙2∙ =2 ∙∙ =2√2≈2.83 (1) NASA/CP—2018-219887 84
Figure 2. Axial load-deflection characteristic and gapping points for hard preloaded bearings  Figure 3. Axial load-deflection characteristic and gapping points for soft preloaded bearings  Soft preloaded bearing pairs have a static and dynamic behavior very different from that of hard preloaded  pairs. Their fundamental difference in stiffness for the two loading directions causes an asymmetric loaddeflection characteristic. Furthermore, off-loaded races may experience considerable axial travel,  independent from the rest of the bearing. Therefore, their motion should be taken into account as additional  mass bodies in a dynamic model. In order to remain concise, the related equations of motion and further  model details are not presented here. As a key effect, when a force equivalent to the preload magnitude is  applied on the preload spring via the adjacent bearing, gapping occurs in the opposite bearing. If the force  keeps increasing, the spring stiffness will dominate the load-deflection characteristic of the bearing system.  Without gapping, the stiffness of the bearing opposite to the spring is dominating. Therefore, the linearized  stiffness of the bearing system can be approximated by the stiffness of a single bearing at the preload  magnitude. The load deflection curve of a soft preloaded bearing pair is presented in Figure 3. Hence, soft -preload0preload Axial deflection-Fgapping-Fpreload0FpreloadFgapping Bearing pair Individual bearing Onset-of-gapping 0preload snubber Axial deflection0Fpreload Bearing pair Bearing pair with snubber Individual bearing Spring Onset-of-gapping NASA/CP—2018-219887 85
on the shaft, it prevents excessive displacements during gapping. The three test units differ from each other  in terms of their preload characteristics, summarized in Table 1.  Figure 5. Cross section of the bearing cartridges (test units)  During the test program, the bearing cartridges were mounted on two different shakers, the first one for low  amplitudes and the second for higher-level vibrations. The test units were only excited along the axial  direction. One accelerometer (two for the stronger shaker) was used for input acceleration control and  monitoring. Another accelerometer was placed on the dummy mass. Furthermore, three load cells were  mounted in the adapter placed between the shaker and the test unit. Their purpose was to monitor the  shaker input force by direct measurement. The phase of the force signals was also used to identify the  direction of the structural modes observed, and to discriminate any non-axial modes. The test set-ups are  depicted in Figure 6. In the left picture, a test unit with standard dummy mass is mounted on the small  shaker. In the right picture, a test unit with additional dummy mass sits on the stronger shaker.  Table 1. Test unit preload parameters  Test ID Preload type Preload magnitude Preload stiffness  3 Soft 20 N 250 N/mm  6 Soft 20 N 900 N/mm  22 Hard 160 N (not applicable)  The test units have been exposed to a number of vibration tests: sine sweep, constant frequency sine with  different amplitudes and random vibration.  Summary of Test Results  For all the tests, the time-domain signals of all sensors were recorded. The sine sweeps were carried out  from 20 Hz to 2 kHz, with an input amplitude varying from 0.1 g to 1 g for the soft preload test units and  from 1 g to 9 g for that with hard preload. The random tests ranged from 0.1 g rms to 9 g rms for all units. The  constant frequency sines were applied at 5 different frequencies around the first main resonance of the test  units (two below, one at the resonance frequency and two above). The amplitude ranged from 0.2 g to 5 g  for the soft preload units and from 1 g to 58 g for the hard preload unit. Most of the tests were repeated  using the additional dummy mass. A comparison of the low-level sine sweeps before and after the highlevel runs confirmed that the dynamic response of the bearing cartridges was not affected by the tests.  NASA/CP—2018-219887 87
Figure 6. Test units mounted on the small and stronger shaker, including instrumentation  Discussion of Test Results  Frequency Domain  A first overview of the results can be obtained by inspecting the estimated Power Spectral Densities (PSDs)  of the output (response) acceleration, for random excitation (see Figure 7). For the hard preload test unit,  the resonance frequency decreases slightly with increasing excitation magnitude. Some other features  have been found in the PSD plots, but a closer analysis of the load cell phases indicated that those features  resulted from a cross-coupling between a radial mode and the axial excitation. For the soft preload cases,  it is found that the resonance frequency (and, hence, the apparent stiffness of the bearing system) starts  decreasing with an increasing excitation magnitude. However, when the excitation becomes high enough  to reach snubber contact, then the resonance frequency of the system starts rising again. Moreover, higher  frequency content is introduced due to contact with the snubber.   Moreover, Figure 8 underlines that the quality factor and the resonance frequency are linked, as functions  of the excitation level. Overall, the observations have confirmed the findings by ESTL derived from the  previous test campaign [1].  Time Domain (for constant frequency sine vibration input)  In the runs with constant frequency sine excitation, the increasing input levels were maintained over  sufficiently long periods of time in order to reach steady state. Thanks to this, the measurement noise and  other random components could be reduced by coherent averaging of the response signals. The averaged  responses are shown for the hard preload case and one soft preload case (#3) in Figure 9. While the hard  preload test unit behaves symmetrically, the acceleration asymmetry is clearly visible for the soft preload  test unit.   At resonance, housing acceleration (input) and dummy mass and shaft acceleration (output) are shifted in  phase. Therefore, the bearings undergo higher deflections. Double integration of input and output  accelerations and subsequent evaluation of their difference allows to estimate the relative displacement  between housing and dummy mass and shaft. Consequently, multiplying the dummy mass acceleration by  the total moving mass (dummy mass plus shaft) results in a good estimate of the dynamic force applied on  the shaft. This allows for a comparison with the theoretical (quasi-static) load-deflection curve presented in  Figure 2. However, it should be kept in mind that the complete equation of motion comprises:  ̈ = ,−{̇ ,̇ , , ,…} (2) NASA/CP—2018-219887 88