ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
The electrical configuration of one module of a 2.5D slip ring consists of 20 power lines designed for 5.4 A  and 20 signal lines designed for 1.1 A. This allows transfer of a total current of about 130 A (forward and  return). For two modules in parallel, the transferred current can be doubled .  All steps performed during assembly have been verified and the process was validated. The final subassembly and assembly pictures are shown in Figure 4.  Figure 4. Sub-assembly and assembly pictures  Four units have then been tested according to the test sequence shown in Figure 5.  Figure 5. Qualification test sequence  The 2.5D slip rings performed all the mechanical and environmental tests without showing any  degradation. Some lines were detected out of specification in terms of dynamic resistance performance  during the environmental tests. After an investigation, it was confirmed that the identified noise was linked  NASA/CP—2018-219887 213
to contamination during integration. It has been assessed that a potential contamination occurred onto the  slip rings rotors while cycling them into the oven. (Rotors were positioned into the oven as described in  Figure 6).  Figure 6. Rotor sketch assembly into the oven  The investigation showed that the tooling jigs the rotor rests onto into the oven have the exact same  shape as the one observed on the rotor. Some Kapton® tape residue was present on the jigs and the  assumption is made that contamination occurred at this stage. Kapton® tape glue has evaporated onto  the rotor track during the temperature cycling. The rotor clearly exhibits a discoloration as depicted in  Figure 7.  Figure 7. Discoloration on rotor track  Right after the tear down activity, a deep cleaning of the rotor was performed to verify any effect this  could have on the resistance final results. Slip ring was re-assembled and tested after the cleaning and  the new dynamic resistance test passed. The difference between the results at the end of the test  sequence and after cleaning is shown in Figure 8 (resistance over one complete revolution on one track).  Figure 8. Noise before cleaning (left) and after (right)  NASA/CP—2018-219887 214
The conclusion has been made that Kapton® glue on the track is the root cause of the deviation we found.  Such contamination can generate noise and resistance test deviations.  The new rotor technology implemented in the 2.5D slip ring has demonstrated its performance in terms  of:  •Insulation properties •Resistance properties •Mechanical properties •Thermal behavior •Wear (GEO and LEO applications) In addition, this design has demonstrated its robustness with regards to the current capacity, by working  without any damage with a current of 6.2 A on all the power lines during 24 hours at hot temperature.  The same level of performance is achieved with this type of technology compared to the standard one,  see Figure 9.  Figure 9. Standard deviation evolution  At the end of the validation test campaign, a tear down was performed on two units: SN001 that  performed the GEO life test, and SN004 that performed the LEO life test (Fig. 10). The tear down  confirmed that the wear present in the slip ring is nominal and that there is still enough gold on the track  after the GEO and LEO life test.  Figure 10. GEO life test (left), LEO life test (right)  In addition, the wear on the wire brushes is as usual and the wire brush preload verification has  demonstrated that 99% of the wire brushes were in the expected range of preload at the end of the life  test.  NASA/CP—2018-219887 215
Conclusion  In conclusion, the 2.5D technology can be used to replace the standard costly and risky concept.  The perspective for further technical evolution is quite large. A great benefit with this kind of design is the  modularity. The number of tracks, the electrical configuration, as well as the number of wire brushes can  be modified to be in line with the customer specification.  Today, there are two technical limitations of this system:  •The stacking of only two modules •The thermal aspect For the stacking, some change in the design can be foreseen to have the possibility to stack an infinity of  modules. The main change will be the output of the wires that shall go out on the external side of the  stator, respectively on the internal side of the rotor.   For the thermal aspect, the result of additional tests has demonstrated that the system can survive  additional current and power, but in a limited range. The temperature at a soldering location is only 10°C  under the allowable temperature. This margin shall be kept for other applications.   From a commercial perspective, the 2.5D PCB technology activity is fully in-line with the development of  new SADM and instrument applications and provides a competitive and technological advantage to our  company. The optimization with an industrialization process is an undisputable advantage against other  space slip ring manufacturers worldwide.  This innovative technology will not only improve the product reliability, the performance and the  repeatability, but will also drastically decrease the procurement lead time. Indeed, with today’s technology  several uncompressible processes are needed leading to a final delivery time of twelve months for the  two first modules. With the 2.5D PCB technology, the lead time will be reduced by at least a factor of two.  Furthermore, the competitiveness at a SADM level being more and more difficult worldwide, the cost of a  slip ring is becoming one of the drivers in the SADM price. With the new 2.5D PCB technology, the final  price of the product would be also reduced.   Finally, an important lesson learned was Kapton® tape with Kapton® glue shall not be used during the  manufacturing, assembly, integration and test of a slip ring.  NASA/CP—2018-219887 216
Quantifying Threaded Fastener Locking  Daniel P. Hess* and Christopher DellaCorte**   Abstract  A mechanism for loosening of threaded fasteners is explained and defined in terms of the self-loosening  moment inherent to threaded fasteners and an external load induced loosening moment. Equations for  these loosening moments are defined. The locking moment needed to prevent loosening due to the selfloosening moment and an external load loosening moment in threaded fastener joints is quantified. This  together with secondary locking moment data provide a basis for specifying secondary locking in  threaded fasteners to prevent loosening. An example is presented for an aerospace fastener.  Introduction and Background  Assembly of a bolted joint generally involves application of a tightening torque to threaded fasteners with  resulting relative angular motion and stretch defined by the thread helix. After assembly, this stretch and  associated potential energy is often held by the inherent friction at the thread and bearing interfaces. This  friction is the primary locking mechanism or feature in preloaded threaded fasteners and is proportional to  preload.  In many cases, a secondary locking feature is added to a threaded fastener [1]. These include  mechanical locking (e.g., cotter pin or safety wire), prevailing torque locking (e.g., deformed thread or  polymer patch), adhesive locking (e.g., anaerobic adhesive), and free spinning locking (e.g., serrated  bearing surface or lock washer).  Use of a secondary locking feature may be required as in some aerospace applications [2] to provide  redundancy and to counter possible loosening and failure in the event of unexpected loads. Often it is  added to correct an observed or documented fastener loosening problem. In cases of fasteners without  preload, it provides the only form of locking in fasteners.  Despite the widespread availability and use of secondary locking features, very little quantitative design  guidelines exist. This paper aims to quantify locking and to help answer the question “What moment does  a secondary locking feature need to provide to prevent loosening?”  Self-Loosening Moment  Threaded fasteners exhibit an inherent self-loosening moment. This self-loosening moment is evident  from the torque equation for a bolted joint. The tightening torque for a threaded fastener in a bolted joint is  generally [3-5] defined as shown in Eq. 1.  2 cost t t p n nr pT F rµµπ β = + +  (1)  Here Tt is the tightening torque, Fp is the preload, p is the thread pitch, µt is the thread interface friction  coefficient, rt is the nominal thread interface radius, µn is the nut interface friction coefficient, and rn is the  *University of South Florida, Tampa, FL ** NASA Glenn Research Center, Cleveland, OH Proceedings of the 44th Aerospace Mechanisms Symposium, NASA Glenn Research Center, May 16-18, 2018  NASA/CP—2018-219887 217
nominal nut interface radius. The first term in Eq. 1 is the torque required to stretch the bolt and the  remaining two terms are the torque required to overcome thread and nut friction.  The removal torque is shown in Eq. 2.  2 cost t r p n nr pT F rµµπ β = − + +  (2)  This removal torque is the torque required to overcome thread and nut friction minus the torque from bolt  stretch. This negative term (Eq. 3) defines the self-loosening moment.  2p self lF pMπ−= (3)  This results from the bolt stretch torque and associated potential energy in the bolt. It is inherent to the  threaded fastener and is proportional to preload and thread pitch.  Loosening Moment from External Loads  In addition to the inherent self-loosening moment in fasteners due to bolt stretch, external loads introduce  loosening moments. An external transverse load on a bolted joint is illustrated in Figure 1.    Since threaded fasteners are made with lateral clearance for assembly, the nut can be moved sideways  by an amount equal to this lateral clearance. Consider the cross section of a bolted joint as shown in  Figure 1 with a force applied to the front of the nut. Since the right side will move with greater difficulty, it  acts as a pivot point about which the nut slips and turns loose. When a force is applied to the back of the  nut as shown in Figure 2, the left side becomes the pivot point about which the nut slips and rotates  loose. The net effect of a dynamic or cyclic transverse force is a ratcheting loosening motion, and the  amount of slip and resulting loosening increases with lateral thread clearance.  Taking the pivot points on the thread pitch diameter and the external transverse load applied at the center  of the fastener, the moment arm is the thread pitch radius rt and the loosening moment from front and  back external transverse loads is shown in Eq. 4.  ext l front t back tM F r F r−= = (4)  Similar loosening moments can result from other external loads such as axial loads, bending loads, and  combined loads.    Primary locking  The friction terms in the tightening and removal torque (Equations (1) and (2)) define the primary locking  moment in a bolted joint as shown in Eq. 5.  cost t primary locking p n nrM F rµµβ = +  (5)  This primary locking moment is dependent on preload and friction. Unfortunately, if sustained cyclic slip  occurs in a bolted joint, this primary locking moment is ineffective and locking must be provided by a  secondary locking feature.  NASA/CP—2018-219887 218
Secondary Locking Moment  In a joint subjected to only an external transverse load, the total loosening moment is shown in Eq. 6.  2p l self l ext l ext trans tF pM M M F rπ− − − = + = + (6)  In cases with sustained cyclic slip where thread and nut friction are ineffective, this total loosening  moment from combined self-loosening moment and external load loosening moment defines the needed  locking moment from a secondary locking feature. Specifically, the needed secondary locking feature  locking moment is shown in Eq. 7.  locking l self l ext lM M M M− − ≥ = + (7)  Prevailing Torque Locking  The tightening torque equation with secondary locking prevailing torque feature added is shown in Eq. 8.  2 cost t t p n n pvr pT F r Tµµπ β = + + +  (8)  Here Tpv is the prevailing torque.  It is independent of preload Fp.  The removal torque equation with secondary locking prevailing torque feature added is shown in Eq. 9.  2 cost t r p n n pvr pT F r Tµµπ β = − + + +  (9)  Even without friction and preload, a locking moment equal to the prevailing torque Tpv remains. Prevailing  torque locking features provide locking even with complete loss of preload. As a result, prevailing torque  locking is often used in applications where complete disassembly and loss of components must not occur.  Well-defined standards exist for fasteners with prevailing torque locking [6-11]. These provide allowable  minimum and maximum prevailing torque for a given thread size.    Secondary Locking Measurement  The locking moment provided by a secondary locking feature can be measured with a torque wrench or  sensor on test specimens or actual hardware. This measurement should be taken as a separate test with  the fastener in a state of zero preload. This is routinely performed in practice for prevailing torque  locknuts, but can also be performed for adhesives, safety lock wire, cotter pins in castle nuts, and other  secondary locking features. Aerospace Fastener Example  As an example, consider an NAS1004 0.25-28 UNJF (approx. M6x1) thread fastener in a joint with a  2400-lb (10.7-kN) preload. The thread pitch is 1/28 inch.  The self-loosening moment is shown in Eq. 10.  2,40013.62 2 (28)p self lF pMπ π−= = = in-lb (1.54 N-m) (10)  NASA/CP—2018-219887 219
References  1. Hess, D, “Threaded fastener secondary locking requirements,” Journal of Failure Analysis and Prevention Joints 17 (2017) 724-730. 2. NASA-STD-5020, Requirements for threaded fastening systems in spaceflight hardware, 2012. 3. Bickford, J, An Introduction to the Design and Behavior of Bolted Joints , CRC Press (1995) 140142. 4. VDI 2230, Systematic Calculation of High Duty Bolted Joints – Joints with One Cylindrical Bolt , Verein Deutscher Ingenieure, 2003. 5. Oberg, E, Jones, F, Horton, H, Ryffel, H, Machinery’s Handbook 25th Edition , Industrial Press, 1996. 6. NASM25027, Nut, self-locking, 250oF, 450oF and 800oF, National Aerospace Standard, 1999. 7. NASM8846, Inserts, screw-thread, helical coil, National Aerospace Standard, 1998. 8. MIL-I-45914A, Insert, screw thread – locked in, key locked, Department of Defense Military Specification, 1991. 9. MIL-I-45932/1C, Insert, screw thread – thin wall, locked in, Department of Defense Military Specification, 1994. 10. NAS3350, Nuts, self-locking, 450oF and 800oF, high quality, National Aerospace Standard, 2006. 11. MIL-DTL-18240F, Fastener element, self-locking, threaded fastener - 250oF maximum, Department of Defense Detail Specification, 1997. 12. Cheatham, C, Acosta, C, Hess, D, “Tests and analysis of secondary locking features in threaded inserts,” Engineering Failure Analysis  16 (2009) 39-57. 13. Hess, D, “Secondary locking feature locking moment measurements,” to appear. NASA/CP—2018-219887 221
where (Application Values)  •αL = Mean Acceleration of the Load •TPPS-0 = Torque at Low Pulse Rate at Motor •FM = Detent Plus Friction at the Motor •N= Gear Ratio •FL = Load Friction •JL = Load Inertia •JM = Motor Rotor Inertia •ηG = Gearbox Efficiency The Torque to Accelerate the Load ( TαL) is calculated by Eq. 2:  TaL=JL×aL (2)  This torque is actually transmitted through the gearbox at every pulse of the system. As a note, since the  gearbox efficiency attenuates the acceleration at the output, it is most conservative to use 100% gearbox  efficiency.   Calculation of Loaded Time and Mean Loaded Velocity Now that we have determined the peak-transmitted torque to the load, we must determine the time the unit  is under load. Stepper motors do not transmit torque between step pulses when the shaft is settled. As long  as the inertia factor is reasonable (under 3.0) the overshoot and stabilization torque are also insignificant.  Therefore, we are primarily interested in the time it takes to accelerate the load at each pulse ( tαL). Which  is estimated in Eq. 3.   taL=2× DQL aL+te (3)  where (Application Values)  •tαL= Time to Accelerate Load •ΔΘ L = Step Size at Load •αL = Acceleration at Load •τe= Motor Electrical Time Constant This results in a time to accelerate the load at each pulse of in a period of seconds. Given the life  requirement of a number of steps a Loaded-Lifetime requirement of can be calculated. There are several  offsetting secondary and tertiary components that affect the actual load characteristics and step kinematics,  however, this analysis is considered conservative and appropriate. System drive electronics, internal  damping characteristics as well as backlash will have minor effects on the step kinematics. These system  variables can be difficult to predict or quantify, and their affects are much less significant compared to the  primary variables defined in Eqs. 1-3.   Now that the torque and time components are established, we must determine the mean velocity when the  mechanical energy is transferred. This calculation is presented in Eq. 4.   wL_Loaded =DQL taL (4)  where (Application Values)  •ωL_Loaded = Mean Loaded Velocity This analysis results in a Mean Loaded Velocity in Radians per second. The analyzed load acceleration  torques ( TαL), Mean Loaded Velocities, and Loaded Lifetime must be used in the gearbox and bearing  NASA/CP—2018-219887 224
•b = proposed constant, for Teflon insulation applications, b = 0.144 The constant in Eq. 7 arises from the success of the chosen insulation system. The equation is used to add  an additional margin of safety to the degradation of electrical components. Introducing ε the exposure ratio,  = (8)  where  •Lo = Electrical Operating Time in Hours •LM = Mission Life in Hours Finally, combing the results into the Brancato Method:  = 100∗( ∗2  10 + (1−)∗2  ) (9)  where  •Lw and T R can be found following Eqn. 5 One alternative to energy methods for calculating insulation life due to radiation is to assume that the  insulation follows the reliability equation. Instead, we propose time be replaced with total ionizing dose  (rads)  =( ). (10)  where  •δ = Total ionizing dose of space radiation during the mission (rad), assumed to be 10000 rads for 5year mission at geosynchronous orbit •η = 4.25E7 rads, proposed constant for insulation reliability, see appendix C Reliability Analysis Toolkit  Using the on-line tools of the Reliability Analysis Toolkit, Ref. [2], we are able to predict the reliability failure  rate of an actuator. The methods used in this approach are derived from the NSWC Handbook of Reliability  Prediction Procedures for Mechanical Equipment. The online tools, using the equations described in 3.3.2  applicable Weibull distribution functions for different components.  Reliability Equations  Using the L 2 to MTBF Conversion, the equations represent a three-parameter Weibull distribution, with δ  shown in the equations; however, for purposes of this tool this parameter (sometimes called "failure free  life") is assumed to be zero, and eliminated from these equations. Note, the variables and symbols used in  this section are consistent with those used in Ref. [2].   Reliability Function R(t): R(t)=e-t næ èçö ø÷b (11)  Probability Density Function f(t): f t( )=b(t)b-1 hbe-t hæ èçö ø÷b  (12)  h(t), instantaneous failure rate: h(t)=f(t) R(t)(13)  NASA/CP—2018-219887 227
The average failure rate is calculated using the Eq. 14:  lAverage =1-R(T) 0TR(t)dt ò (14)  System Input Parameters:   1. Shape parameter ( β): Weibull shape “Beta” factors. From Ref. [2]. Usages of Typical or High-End factors are provided. A Beta factor of 1.2 is used for windings, 1.3 for Ball Bearings and 2.0 for Gearboxes. Higher complexity assemblies utilize higher β factors. 2. Characteristic Life ( η) is the 63.2% failure point for a mechanical system. The Characteristic Life is determined by the L 2 life and the β factor. For instance, for a β of 1.3, as for a bearing, the Characteristic Life is about 20x the L 2 life. For a β factor of 2.0, the multiplication factor is about 7x the L 2 life. η can be calculated exactly by using Eq. 15. =−1−  (15)  3. Maintenance interval for item renewal (T): No maintenance of this product is required or assumed. A value of 100 years is used in the on-line tool equations. MTBF with the Reliability Toolkit  The on-line Reliability Toolkit was used to generate a MTBF, from the β function and the L 2 predicted life.  The same equations used in the on-line tool are available in this paper. The lifetime estimations of the  bearings and gears are based on conservative de-ratings of fatigue and wear-out considerations, with  margin, the failure rate of these estimations will be proportional to the required life over the predicted life.  This method has been used in reliability analysis presented to, and accepted by many other programs and  subject matter experts in the industry. The justification for this analysis methodology is the additional  manufacturing precision, workmanship testing and quality inspections that take place for space flight  hardware to eliminate infant mortality and other workmanship issues that affect commercial hardware.   MTBF can be calculated using Eq. 16  	 =  ,     (16)  where, γ(a,x) is the lower incomplete gamma function.  Equation 11 can be ‘altered’ by setting MTBF = T. By doing so, a ‘Maintenance Free’ interval is achieved  i.e. the part on average would fail at the maintenance interval. The equation reduces to η/β*Γ(1/β) for ‘No Maintenance. See Appendix C for the method and code. Redundancy Factor for Windings  Since there is selectable redundancy for the motor and resolver windings, we may use the Redundancy  Factor described in the Reliability Toolkit. We use the Effective Failure Rate on One Standby Offline Unit  with One Active On-line Unit Required for Success (without repair), See Eq. 20.  Reliability Number  We essentially have two separate life analyses to consider, the mechanical reliability and the electrical  reliability. The windings and the electrical insulation system are exposed to radiation the entire time in orbit,  so regardless of powered-on time we assume the time on orbit degrades the insulation system. For a fiveyear orbital mission for a high-radiation, geosynchronous application, we have an electrical life requirement  NASA/CP—2018-219887 228
(tE) of 44,000 hours of operation and 2E+04 Rads Total Ionizing Dose (TID) of radiation during mission life.  Refer to Appendix A for a table of analysis parameters.   To calculate the Mechanical Reliability for the operational profile from the MTBF, Eq. 17 is used.  RtM=e-tM MTBFMé ëêù ûú (17)  where  •RtM = Mechanical Reliability •tM = Time of loaded operation (Hours) •MTBF M = Mechanical Mean Time Between Failures (1.20E+05) This results in an R tM = 0.99736  Similarly, the Electrical Reliability is calculated by:  RtE=e-tE MTBFEé ëêù ûú (18)  where  •RtE = Electrical Reliability •tE = Time of exposure, in Hours (44,000) •MTBF E= Electrical Mean Time Between Failures using (8.95E+12) calculated using γ de-rated Averaged Brancato method. Which results in an R tE =0.99999  The insulation reliability is calculated in equation (19)  = (19)  where  •RtM = Insulation Reliability •δ = Total Ionizing Dose (10000 rads) •MTBF  = Insulation Mean Time Between Failures (4E+7) Which results in an R i =.99975.  And the overall reliability ( RO) is calculated by:  =∙∙ (20)  Therefore, the overall reliability Ro = 0.99701 is achieved.   The MTBF for n redundant units, 1 spare with no repair interval:  =∗∗(+ 1) (21)  where  •n = number of active units •P the probability that the spare will work The probability of the spare working can be treated as that units’ reliability. Using these equations for both  cases of the electrical components results in  Ro = .99736 for the  factored electrical decay. See  Appendices A and B for more information on reliability values. If we needed to increase the resultant overall  reliability, we would address the two lowest MTBF components. In this case, the Mid-Speed Carrier  NASA/CP—2018-219887 229