ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
Bearings that see the radial and thrust loads from the bevel mesh and the low-speed planetary module are  the reliability-driving components. Possible methods to increase reliability in these components is to  increase the gearbox and bearing size, or possibly change materials to increase the L 2 ratings.   Conclusion  Requirements for torque margin may have structural and endurance implications for transmitted torque  through the mechanism that must be considered. Non-current limited applications may also have significant  torque consequences at extreme conditions that are often not analysed. The paper also details a  methodology for calculating reliability using approved analysis techniques and methodologies. We also  introduce two types of analysis for de-rating life estimates for motor insulation systems, due to insulation  aging effects due to prolonged exposure to radiation.   References  1. “Primer – Stepper Motor Nomenclature, Definition, Performance and Recommended Test Methods” Starin et al. Proceedings of the 42nd Aerospace Mechanisms Symposium, NASA Goddard Space Flight Center, May 14-16, 2014 2.http://reliabilityanalyticstoolkit.appspot.com / 3. “Estimation of Lifetime Expectations of Motors” Emanuel L Brancato, iEEE Electrical Insulation Magazine, May/June 1992-Vol. 8 No.3 4. “Analyze of Reliability of Gears” Dr. N. Ungureanu et al 5. “How to Determine the MTBF of Gearboxes” Dr. Gerhard G. Antony Neugart, Paper for 2007 AGMA FTM 6. NSWC Handbook of Reliability Prediction Procedures for Mechanical Equipment, May 2011. 7. “Nuclear and Space Radiation Effects on Materials”, NASA, July 1970. 8. “Design and Development Implication of High Inertia Earth-Science Mechanism Actuator”, ESMATS 2017 Appendix A – MTBF and Reliability Table - Mechanical  A B C D E F G  Subsystem Module or  Component Beta  (Weibull  Shape  Parameter) L2 Life  (Hours) Characteristic  Life (Hours) MTBF  (Hours) Average  Failure Rate  (Hours) Reliability  N/A β See 5.2 η Ref. [2] λ Eq. 5  Motor Rotor Bearing 1.3 4.50E+05 9.05E+06 9.57E+06 1.05E-07 0.99997  High Speed Planetary Module 2 1.00E+06 7.04E+06 7.79E+06 1.28E-07 0.99996  High Speed Planet Gear  Bearings 1.3 1.00E+05 2.01E+06 2.13E+06 4.70E-07 0.99985  High Speed Carrier Bearings 1.3 5.27E+05 1.06E+07 1.12E+07 8.93E-08 0.99997  Mid Speed Planetary Module 2 1.00E+06 7.07E+06 7.79E+06 1.28E-07 0.99996  Mid Speed Planet Gear  Bearings 1.3 8.75E+05 1.76E+07 1.86E+07 5.38E-08 0.99998  Mid Speed Carrier Bearing 1.3 2.94E+04 5.91E+06 6.25E+05 1.60E-06 0.99949  Right Angle Bevel Gears 2 1.00E+06 7.04E+06 7.79E+06 1.28E-07 0.99996  Right Angle Carrier Bearing 1.3 3.17E+07 6.38E+08 6.74E+08 1.48E-09 0.99999  Right Angle Support Bearing 1.3 5.40E+06 1.09E+08 1.15E+08 8.71E-09 0.99999  Low Speed Planetary Gearing 2 2.50E+04 1.76E+05 1.95E+05 5.14E-06 0.99837  Low Speed Planet Gear  Bearings 1.3 1.03E+05 2.07E+06 2.19E+06 4.57E-07 0.99985  Low Speed Carrier Bearings 1.3 2.34E+06 4.71E+07 4.97E+07 2.01E-08 0.99999  Total 1.20E+05 8.33E-06 0.99736  NASA/CP—2018-219887 230
Appendix B – MTBF and Reliability Table - Electrical  A B C D E G H I  Subsystem Module or  Component Beta  (Weibull  Shape  Parameter) L2 Life  (Hours) Characteristic  Life (Hours) MTBF  (Hours) Reliability R o Redundan t Ro  .1Avior standard* at  te=44000 1.2 2048000 5.29E+07 2.76E+07 0.99841 0.99577 0.99656  .1 Averaged Brancato  at te=44000 1.2 2.60E+09 2.69E+10 1.40E+10 0.99999 0.99736 0.99736  ψ at te=44000 1.2 1.47E+11 3.80E+12 1.98E+12 0.99999 0.99736 0.99736  γ Averaged Brancato  at te=44000 1.2 6.63E+11 1.13E+13 8.95E+12 0.99999 0.99736 0.99736  Calculations of reliability and MTBF include both the On-Axis Resolver and Motor Windings  *The Avior “old” standard is the Brancato L 2 life at 120˚C multiplied by .1 The averaged Brancato method can be found in Eq. 9 and in Appendix D A value of -30˚C was used for the cold or non-powered state of electrical components A value of +120˚C was used for the hotspot temperature of electrical components Appendix C  The following code was written in Anaconda Spyder to calculate η, MTBF, λ, and no maintenance interval (T)  # denotes code comment  from scipy.special import gammaincc, gamma  from math import *  from math import exp as e  B = β  #beta shape parameter  L2 =L 2 #L2 life  n = L2*(-log(.98))**(-1/B) #η, characteristic life for L2  le = 5*365*24 #mission life for electrical components (Hours)  lm = 318  #usage life requirement for   T = 50   #initial guess in hours of no maintenance interval  R = e(-(T/n)**B) #reliability function  MTBF = (n/B*gamma(1/B)*gammainc(1/B,(T/n)**B))/(1-R) #see ( ◊)  err = abs(T-MTBF)/MTBF*100 #percent error  ep = .1   #final percent error  while err>ep: # subroutine solves T for conditions stated after Eq. 11 by simply convergence   T = MTBF # Iterated map on T   R = e(-(T/n)**B)   MTBF = (n/B*gamma(1/B)*gammainc(1/B,(T/n)**B))/(1-R)   err = abs(T-MTBF)/MTBF*100  print(n,MTBF,1/MTBF,e (-lm/MTBF)) #reports results η, ΜΤΒF, λ, and reliability R n  print(T/365/24,"Years") #no maintenance life in years ◊scipy gammainc is normalized, so it has to be multiplied by gamma from scipy.special import gamma B = β  #beta shape parameter  L2 =L 2 #L2 life  n = L2*(-log(.98))**(-1/B) #η, characteristic life for L2  R = e(-(T/n)**B) #reliability function  MTBF = n/B*gamma(1/B)  NASA/CP—2018-219887 231
Development of the NEA Mini for Low Load Applications  John Sudick∗ and Geoff Kaczynski∗ Abstract  Small satellites require simple, reliable and low-cost release devices. Nichrome burn wire solutions have  been embraced by the Small Satellite community because they require low power, are inexpensive and  readily available. Industry feedback gathered over the past 7 years at the Small Satellite Symposium in  Logan, Utah, suggests that the failure rate of Nichrome wire is too high and a more reliable solution is  needed. Based upon this industry feedback, NEA started working to see if we could adapt our GEO NEA  battery bypass switch into a release mechanism for small sat applications.   Looking across our stable of qualified release products, NEA  determined that the release mechanism utilized in our battery bypass  switches are small, light, low power, low shock, and have high  reliability. Our goal for this paper was to perform a series of load and  shock output test to determine if the release mechanism in our  bypass switch could be modified for used in standard release  mechanism applications.   Introduction  The small satellite market has expanded quickly over the last 5 years. With affordable access to space  becoming a reality, the need for a 50 - 200 lbf (89 – 890 N) class, low cost, low power, low shock, high  reliability release mechanism has become a reality. Customers are now looking for ways to increase their  mission success rates without breaking the bank. NEA’s goal was to determine if we could leverage the  volume of our battery bypass mechanism product line to offer a low cost high reliability release mechanism.  The development team’s charter stipulated that they had to utilize off-the-shelf common stock parts that are  utilized in our bypass switch release mechanism. The stock parts could be modified for the application, but  no new parts were allowed. This paper walks through the initial test data and lessons learned during the  development of the NEA Mini 9300 release mechanism.   * NEA Electronics Inc., Moorpark, CA Proceedings of the 44th Aerospace Mechanisms Symposium, NASA Glenn Research Center, May 16-18, 2018 NASA/CP—2018-219887 233
Basic Design  The development of the NEA Mini centered on being able to utilize the common release mechanism from  our battery bypass switches. The release mechanism in the bypass switch (Figure 1) is similar in operation  to our standard release mechanisms and utilizes the same patented fuse wire technology.   Figure 1. Battery bypass switch  For switch activation, an electrical pulse of 1.2 amps or greater is sent to the mechanism. The fuse wire  breaks and releases the restraining wire that holds two spool halves together. A preloaded spring assembly  pulls a plunger forward once the spools are released. The shock is minimized by releasing the stored strain  energy over a longer period of time. The electrical characteristics for the fuse wire assembly are time and  current dependent as shown in Figure 2:  Figure 2. NEA Mechanism Actuation Curve  Cover removed for clarity  NASA/CP—2018-219887 234
Based on qualification, NEA’s largest bypass switch spring imparts a load of 65 lbf (289 N) on the spools  of the release mechanism, so the team knew that the release part of the device could hold a least 65 lbf  (289 N) of load. The next questions to be answered were, how high of a load could it handle and what were  the limiting factors in the load capability of the mechanism? Was the limiting design factor the Spool  Material, Restraining Wire, Fuse Wire, or Release Rod? NEA started by load testing the bypass switch  release mechanism to determine its maximum load capability. The team designed tooling to push on spools  until the unit failed. The load released at approximately 750 lbf (3300 N) with the restraining wire breaking  first (Figure 3).   Figure 3. NEA Ultimate Load Test  Now that we knew the ultimate load of the bypass switch release mechanism, we next turned our attention  to the design of the release rod. As discussed above, to hit our cost targets, the team was only allowed to  alter stock bypass release mechanism parts. They reviewed the piece part detailed drawings to determine  our design constrant for the size of release rod. It was determined that the size of the hole that could be  drilled between the guide pins that restrain the the spools was the limiting factor. This is shown in Figure 5.  Our calculations showed that we could install a hole in the assembly that would support the release of a #1  fastener (release rod) and that this rod would have the structural capability to hold 450 lbf (2000 N).   We then looked at the contact pressure at the conical interface between the release rod and the spools. As  it turned out, this was the limiting factor relative to the maximum load capability of the device. The spools  are machined from a high-strength engineered plastic that has a maximum compressive strength of   28,900 psi (199 MPa).   Figure 4. Release Rod Load Calculations  NASA/CP—2018-219887 235
The design team armed with the ultimate load data and release rod load data set off to do our first prototype  test. We modified a bypass release mechanism and machined the release rod in our rapid development  lab. The release rod was made by modifying a commercial #1 socket head cap screw. Figure 5. Release Rod and Release Mechanism Modified  At this point we had our first lessons learned. Our first experiment (Ultimate Load) along with the release  rod load calculations theorized that the load capability of the mechanism should be near 450 lbf (2000 N).  The first test of the mechanism produced contrary results. The initial load test done with the bypass switch  mechanism was performed by pushing on the spools (using an Instron test apparatus) with a constant load  rate; hence, load creep or loss was not measured. In reality, the plastic spools of the baseline mechanism  deflect radially outward at applied loads greater than 150 lbf (660 N). While this test was considered  acceptable because the load was within the desired 50 to 200 lbf (89 – 890 N) load range, the team felt  they could improve the results.    A detailed finite element analysis confirmed that the spools were deflecting only in the area above the  restraining wire; that is, in the area unsupported by the restraining wire. Based on this analysis, the team  elected to modify the spools by lowering the conical section deeper within the spool, assuring that the radial  load generated by the conical area was fully supported by the restraining wire. This deviated from the  charter of the project because parts needed be altered such that stock parts could not be used. The results  of the second test were much more favorable and improved the load capability of the mechanism to more  than 300 lbf (1330 N).   Limiting Design Factor  NASA/CP—2018-219887 236
Figure 6. Load Creep Test Results  Now that load characterization was complete, the team moved on to capture shock output data. Given the  small size of the device, a modest 12 inch by 12 inch (30 cm by 30 cm) aluminum plate (1/8-inch (3-mm)  thick) was used for shock output testing. The test set up is shown in Figure 6.  Figure 7. Shock Test Setup 200205210215220225 -1 2 5 8 11 14 17 20 23 26 29 32 35Load (lbf) Time (hrs)Load Creep Test Results (typical) NASA/CP—2018-219887 237
NEA built and tested 5 (five) units for the shock output test. The units were loaded to 250 lbf (1100 N) and  then actuated with 4.5 amps of current. The results of the output test are shown in Figure 8.  Figure 8. Shock Output Test Results  Conclusions and Lessons Learned  1. NEA produced a 100-lbf (440-N) release mechanism using common parts with our production bypass switch mechanism. 2. Much higher loads (>250 lbf (1100 N)) are achievable with slight design modifications to common parts. 3. By changing from plastic spools to metal spools, loads of 450 lbf (2000 N) can be achieved, but at a higher cost. 4. The shock output of the mechanism is well below expectations and industry standards. 5. More attention to the rod to spool interface could have saved a design iteration.100 1000 100000.010.101.0010.00100.001000.00 Frequency (Hz)SRS ShockSRS Response, 5% Damping 1X 2Y 3Z 4X 5Y 6Z NASA/CP—2018-219887 238
of a 2-DoF linear stage demonstrator and continued with a second chapter consisting in the development  of an end-to-end production strategy addressing the optimization of the material properties. The main  conclusions derived from the characterization of these samples were that - in their optimized version -  their tensile properties were similar to those of the commercial grade alloy, while their alternate bending  fatigue lifetime was greater than 15 million cycles under cyclic loads near to 50% of their Yield Strength.  Based on these encouraging results, a third chapter was opened in 2017 with the aim to design, produce  and test several topologically optimized parts and mechanisms.  Figure 1. AM-SLM produced 316L stainless steel 2-DoF linear stage. Flexures thickness = 300 µm,  aspect ratio respectively 80 and 123 for the blades of the double and simple parallelograms  Materials Selection, Development and Validation  The material choice was oriented toward high-strength stainless steel alloys available in powder form,  with the aim of approaching the exceptional mechanical properties and stress corrosion cracking (SCC)  resistance of MARVAL X12, the material usually chosen by CSEM for demanding applications. This highstrength precipitation-hardened stainless steel offers both high SCC resistance [2] and high fatigue  resistance when it is submitted to alternate bending deformation, a parameter which was experimentally  verified through several internal fatigue test campaigns. The alloy chosen – Concept Laser’s CL92PH – is  an equivalent of the widely used and studied 17 ‐4PH martensitic precipitation-hardening stainless steel  [3, 4]. Recently, the AM process for titanium alloy has been optimized following the same procedure.  Other prototypes have been made of aluminium AlSi10Mg and copper by SLM.  SLM Process Optimization  The optimization of the SLM parameters was performed in an iterative manner with the aim to minimize  porosity and surface roughness, whilst seeking to optimize micro-structure for high mechanical  performances.   Figure 2. Fatigue sample cross-sections before and after SLM parameters optimization.  The thickness of the blade is 350 µm.  NASA/CP—2018-219887 240
Among the process parameters evaluated, such as layer thickness, laser beam power, focus point or  scan speed, the laser pattern was confirmed to be the key parameter leading to homogenous material  quality on both massive and thin geometries, as shown in Figure 2.   The impact of the manufacturing direction was also investigated. The only notable differences were found  on the transverse cut where Y samples showed a finer micro-structure compared to X. These visual  differences were all suppressed after the Hot Isostatic Pressing (HIP) treatment, which removed all  porosity and improved the micro-structure, such that X and Y could no longer be distinguished (see  Figure 3). The analysis confirmed that the HIP treatment permits the removal of the residual porosity and  the improvement of the microstructure in terms of homogeneity and grain size.  Figure 3. Sample cross-sections before and after HIP processing.  Tensile Test Results  The tensile tests were conducted for two groups of samples representing two plausible final heat  conditions for mechanical parts of a compliant mechanism. The samples were produced according to the  three manufacturing directions, X-Y-Z, to highlight potential anisotropies in the final heat condition states.  The second objective was to quantify the difference expected between HIP and non-HIP material. The  tensile test samples were designed according to the ASTM E8/E8M − 15a standard. Raw rods were  produced and submitted to the heat treatment sequences. Finally, the rods were machined to their final  geometry and tested.  The results show that for both heat treatment conditions, the SLM produced samples show similar or  higher Rm and Rp 0.2 compared to the commercial grade 17-4PH. The elongation at break for non-HIP  samples tend to show a fragile behavior which is confirmed by the fracture inspection. The tensile tests  revealed minor anisotropy according to the manufacturing directions, X showing the highest Rm & Rp0.2 ,  followed by Z and Y. The overall relative variations of Rm & Rp0.2  versus the manufacturing directions  are below 5% for the HIP samples and 10% for non-HIP samples, which tends to show a slight  improvement on anisotropy induced by the HIP treatment.  Fatigue Test Campaign  The graph of Figure 4 compares the approximated S-N curves and fatigue limit estimations Sf of each  fatigue test sample (FTS) group. As expected, the group C whose FTS were manufactured from  commercial grade 17-4PH and machined according to the best WEDM-based protocol shows the highest  fatigue performance with Sf C estimated at 680 MPa. This group is to be compared with group B, where  FTS were machined according to the same WEDM-based protocol, but from additively manufactured and  HIP raw material. The subsequent loss in fatigue performance is around 25% with Sf B estimated at  510 MPa. NASA/CP—2018-219887 241
Figure 4. S-N curves estimates for SLM (CL92PH) and WEDM (17-4PH) samples. The fatigue limit of  MARVALX12 was determined experimentally during the MTG-CCM test campaign.  Design for Additive Manufacturing  Using AM technologies to reproduce a part whose design was driven by a conventional manufacturing  strategy such as turning and milling is often not pertinent, since it does not take into account most of the  advantages given by AM. For example, the complexity of the assembly can be reduced by designing a  monolithic mechanism or by decreasing the number of parts and combining functions.  To ensure an optimized redesign for AM, a system engineering vision will allow understanding all  parameters pertaining to the mechanism and therefore being aware of all key requirements to be  considered during the design phase. On the other hand, the design constraints associated with AM are to  be well understood.  A major limitation of the current AM technologies (e.g. SLM) is the need of support structures when the  part comprises overhanging areas with an angle of usually less than 45 degrees with respect to the  horizontal plane. Therefore, keeping in mind the following strategies during the redesign will contribute to  obtain better results and will ease the removal of the part from the baseplate:  •redesign with overhanging angles greater than 45 degrees from horizontal •changing the part orientation to minimize support structures •allowing support material in dedicated areas Examples of parts and compliant structures redesigned for Additive Manufacturing  A successful product redesign for AM is the rotor of a slipring  performed in partnership with RUAG Space Switzerland Nyon, where  the support material could be avoided. The production of complex  geometries by AM-SLM allowed us to integrate the electrical  conductors within the rotor structure leading to a drastic reduction in  the number of parts – from 36 to 1 for a 12-track rotor – and therefore a  significant reduction of the manufacturing and assembly time. Figure 5. Slipring rotor made  by Additive Manufacturing  NASA/CP—2018-219887 242
For compliant mechanisms, the redesign is more challenging since the design guidelines of most of the  AM processes recommend avoiding thin structures and abrupt thickness variations, which are  characteristic of flexible structures. CSEM's approach is therefore to work on several aspects of AM in  parallel, such as redesign, thin structures manufacturing and testing, mechanical and thermal postprocesses, Finite Element Modelling (FEM) and topology/parametric optimization.   Topology Optimization is actually the most efficient approach to improve various properties of structural  parts designed for AM, such as mass reduction, eigen-frequency tuning, thermal transfer, thin-wall  thickness optimization, etc. Unfortunately, the use of commercial software is currently limited to the  optimization of structural geometries, since the algorithms are not able to handle flexible structures. Due  to their own nature, the algorithms try to decrease  the compliance of the structure (see example in  Figure 6), which is the opposite of what is requested for flexible structures. To optimize those, we need to  determine where and how to locally increase the compliance in order to provide the requested movement.  Since commercial software solutions were not able to fulfil this need, some particular tools have been  developed internally to be able to generate the geometry for compliant mechanisms.  Figure 6. Example of CLUPI (EXO MARS Mission) bracket redesigned with topology optimization.  Left: classical bracket, middle: geometry optimized to maximize the stiffness with a given mass  diminution, right: part additively manufactured in titanium alloy.  An intermediate approach currently under investigation at CSEM is to improve the rigid parts of the  mechanism, keeping conventional flexure geometries as they have been used successfully up to now.  This concept is currently under investigation at CSEM with the redesign of first, the HAFHA pivot alone,  before integrating the result in a more complex scanning mechanism using two such pivots. Some  pictures of the preliminary concepts are shown in Figure 7 and 8.  Figure 7. Example of HAFHA pivot redesigned for monolithic AM with a given print direction.   Left: classical pivot, middle: optimized for AM with support bridges in red, right: AM pivot after  bridge removal. NASA/CP—2018-219887 243