ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
358 Design improvements for the third generation latching system    The load capability has been significantly increase d while the overall mass of the latching system  could be slightly reduced (with respect to the la tch/latch tab concept, see Figure 6). At the same  time, even more standard materials have been applied, like Al 7075 for the latch housing.   A substantial improvement in mechanical effi ciency and reliability could be realized through a  simplified design that reduced the number of linkages in the load path of the latch primary drive to  a minimum. The short drive train to the latch hook also leads to accurate control in the  synchronization of the release of all latches.    The latch design includes a secondary release mechanism, which:   o accomplishes the dual fault tolerance requirement.    o takes into account the Limited Life Item policy for pyrotechnics (i.e., the pyrotechnic release  capability being provided on t he visiting vehicle only).  The latch secondary drive consists out of tilti ng the latch hook by rupturing the hold-down bolt.  Rupturing the bolt can be done with a Frangibolt actuator  or with a rich cut bolt (pyrotechnic).        The joints in the direct load path of the latc h are now equipped with roller bearings instead of  bushings, resulting in an increased drive efficiency.    The load path from the interface load to the load limiter has been shortened, resulting in minimal  hysteresis in the load limiter operation.     

       Figure 6  Comparison of 1 st generation structural latch vs. 3rd generation structural latch 

    Design load: 34 kN (7640 lbf)   Single fault tolerant   Mechanical drive efficiency: 66%     Design load: 57 kN (12800 lbf)   Dual fault tolerant   Mechanical drive efficiency: 80%    Reduced mounting envelope 1st generation structural latch 3rd generation   structural latch  NASA/CP-2010-216272
359 Design verification tests on the third generation latching system   Two prototype hardware models are built and they have recently been functionally and structurally tested.  By test, a few less-favorable consequences of this generic simplified latch design have been verified for their potential impact on the functionality of the latch.   The investigated points of concern of the pr esent latch design were the following:     Due to the latch primary drive kinematics consisting out of a single 4-bar linkage, the latch hook  will no longer land on its mating counterpart in a stra ight vertical trajectory. Instead, the hook now  has to slide off the mated hook for release. Desp ite the fact that the latch hooks have been Niflorcoated, friction between the contact surfaces has been increased compared to the separation  scenario with a roller interface (being used in the first generation latching system).    Release capability tests with the push-off force 1 have been executed. The push-off force being  applied to the latch is the worst-case condition of two side-by-side latch failures out of all 12  latches.  Test conclusion: all tests did demonstrate the capability of the hooks to release without a  problem.    A preloaded compression spring to the side of the latch hook has to force the hook to close when  rotating the latch crank to the closed position.  Test conclusion: test did show that this func tionality was also working fine while an LVDT  attached to the other side of the latch ho ok monitors the movement of the hook.  

Conclusion  

Testing of the latest latch system has demonstrated the capability to design a stable structural connection  device that is made dual fault tolerant for both doc king and undocking of the IBDM. Design optimization of  the mechanism has been realized through analysis and validation of three concept generations. This was accomplished by consistently taken into account the numerous lessons lear ned that have been gathered  along the way. It resulted in a generic simplified design  of the IBDM latching system that features an  optimized mechanical drive efficiency 2, mass efficiency3 and reliability4.                                                        1 A push-off force is being exerted on the latch by an energized separation system  that must provide the  IBDM with a prescribed delta-velocity in case of an expedited departure.    2 Mechanical drive efficiency being expressed as t he ratio between the required drive torque without and  with friction losses in the linkages.   3 Mass efficiency being expressed as the ratio for the load capability versus the mass of the latching  system.    4 Reliability being expressed as the inverse of the amount of moving parts per drive (considering each  moving part as a potential single point of failure).   NASA/CP-2010-216272
361 Development of the Upgraded DC Brush Gear Motor for Spacebus Platforms 

Robert H. Berning III* and Olivier Viout** 

Abstract   The obsolescence of materials and processes used in the manufacture of traditional DC brush gear motors has necessitated the development of an upgraded DC brush gear motor (UBGM). The current traditional DC brush gear motor (BGM) design was evaluated using Six-Sigma process to identify  potential design and production process improvements. The development effort resulted in a qualified UBGM design which improved manufacturability and reduced production costs. Using Six-Sigma processes and incorporating lessons learned duri ng the development process also improved motor  performance for UBGM making it a more viable opti on for future use as a deployment mechanism in  space flight applications.   Introduction   DC brush gear motors have been used for several years in various spaceflight applications because of their many favorable design features. They are extr emely efficient at converting electrical energy into  mechanical energy using only simple control electr onics. Existing qualified DC brush gear motors for  space flight applications however, use some obsole te materials and processes in their design and  construction. The intent of this development was to review the existing BGM design using the Six-Sigma process to identify potential design improvements and to  select replacements for the obsolete materials  and processes. This paper documents the development and qualification of a UBGM for use as a solar array deployment mechanism on t he Spacebus satellite platform that  maximizes motor performance,  lowers overall drag, and optimizes manufacturability.  The BGM has to operate in hostile environmental conditi ons during test and flight. The proper selection of  materials is critical. Factors that must be considered include:  1. Operate in ambient air, up to 55% relative humidity. 2. Survive random vibration (32.3 G rms) 3. Survive in vacuum (1.0 x 10 -5 torr)  4. Operate in vacuum (1.0 x 10-5 torr) from -50°C to +80°C.   5. Survive in vacuum (1.0 x 10-5 torr) from -50°C to +125°C. 

Background 

The qualified BGM shown in Figure 1 consisted of a DC brush motor and a multi-stage planetary gear train. The design used brush material that was since discontinued, some obsolete materials, and  employed non-forgiving process-driven steps that re sulted in high manufacturing costs. The redesign  addresses materials and processes, manufacturing changes, and test tooling improvements that are  necessary for future successful production of the new upgraded DC brush gear motors.                                                        * Moog Chatsworth Operations, Chatsworth, CA USA  ** Thales Alenia Space, Cannes la Bocca Cedex, France 

Proceedings of the 40th Aerospace Mechanisms Symposium, NASA Kennedy Space Center, May 12-14, 2010  NASA/CP-2010-216272
362 Purpose of Redesign:   1. Enhanced producibility  2. Improved functional pe rformance characteristics  3. Reduced delivery schedules  4. Increased robustness  

Figure 1. Existing BGM  

Development   The existing BGM design was analyzed and a 3D CAD model was created in Unigraphics. Prior failures and manufacturing problems were reviewed for area s of improvement. A Six-Sigma product assurance  process was conducted. Trade studies were performed on major assemblies and a detailed tolerance analysis was completed to identify potential interferences.    A Six-Sigma process improvement team was established. Process walk-throughs were completed on six assemblies and three piece parts from the existi ng manufacturing and build cycles. Personnel were  interviewed and fabrication, assembly and test proc esses of the existing BGM units were observed. Forty  eight items for improvement were identified. Trade studies were initiated on all subassemblies and major components. Design and manufacturing process changes were completed to address all identified issues. The following major areas of potent ial improvement were identified:  Process Improvements  Commutator soldering and inspection Armature paint integrity Armature insulation  Performance Improvements  Optimized motor speed and motor torque Predictable gear head drag  NASA/CP-2010-216272
363 A gear head trade study was completed to develop a cons istently producible design with predictable gear  drag over the required temperature range. Review of the existing gear head design and a detailed  tolerance analysis showed a potential interference at cold temperatures, high drag in the first & second stage bushings, a material combination prone to galling (same gear material used on mating gear teeth) and a high sensitivity to gear center distance shift. The following trade study criteria were selected for  gear head design improvement:  Provide similar gear ratio Non-binding operation at extreme temperatures Manageable internal loss Robust design Non-galling material combinations   Three different gear head concepts were selected for design and testing: a completely redesigned gear  head (option #1), a harmonic drive gear head (option #2), and a modified existing gear head using radial  ball bearings (option #3). Engineering models of all options were fabricated and tested. The redesigned  gear head had higher and inconsistent drag over t he required temperature range. The harmonic drive  gear head exhibited significantly hi gher drag at ambient temperatures, so no further testing was required.  The modified gear head using radial ball bearings (opt ion #3) was ultimately selected based on its low  and consistent drag over the required temperature range. Table 1 lists gear head drag of engineering  models over the required temperatur e range. The modified gear head does not exhibit interference over  the required temperature range, has reduced drag in the first and second stages, has no galling material  combinations and uses a one piece ring gear to minimi ze sensitivity to gear center distance shift.    

A new brush material was identified and selected at  the conclusion of the motor trade study. Detailed  review of the existing motor design revealed inefficient processes, high brush drag and use of discontinued brush material.    The brush assembly consists of a carbon composite br ush, shunt wire, cap, and spring. Eight different  brush materials were considered and four were selected for testing. All brushes were tested for motor performance, resistance, drag, spring force, brush wear, commutator wear, smearing, debris, and manufacturing yield. Brush material option #2 and #3 were eliminated due to low motor torque. Brush material option #1 was selected due to poor performance of option #4 in vacuum. Table 2 lists development brush performance. The se lected brush material is softer than the existing brush material,  resulting in higher motor torque, lower brush drag and less commutator wear.         Description -50°C +23°C +80°C -50°C +23°C +80°C SN042 Existing Design 0.22 0.06 0.07 0.31 0.09 0.10 Option #1 Redesign 1.78 0.60 0.44 2.52 0.85 0.63 Option #2 Harmonic drive 3.53 5.00 Option #3 Radial Bearing 0.16 0.01 0.02 0.23 0.02 0.03Gear Drag @ 450 RPM (N-cm)Gear Drag @ 450 RPM (in-oz)Table 1.  Gear Head Drag NASA/CP-2010-216272
364 Table 1. Brush Performance 

The motor trade study considered all assemblies a nd machined parts. The prod ucibility of existing motor  is poor due to the need for frequent rework resulting in high production costs. Sta ck fabrication, coating,  and attachment methodology were upgraded to current Moog procedures. All uncontrollable and  unnecessary processes were replaced or eliminat ed. For instance, existing BGM commutators are  machined after final armature assembly putting th e completed armature at risk. UBGM commutator  processing was moved to the piece part level to lower the risk to hardware. The soldering process was updated to the current standard. Table 3 shows increased motor torque with new brush materials and design and manufacturing changes. 

Table 2 Motor Performance 

 The overall development of the upgraded brush gear mo tor was successful. All issues discovered during  the Six-Sigma process were addressed. After develop ment was completed a qualification unit shown in  Figure 2 was fabricated to production paper work, us ing production processes and tooling. The unit was  subjected to qualification testing which included vibrat ion, thermal vacuum exposures and life tests. The  qualification unit successfully passed all qualification and life tests with no findings.    

Figure 2. UBGM Motor Torque Commutator Brush  Material SN 0042 N-cm (in-oz)Wear in AtmosphereWear in VacuumDrag  N-cm (in-oz) Debris SmearingResistance (Ω) Yield Wear 1 1.77 (2.50) Good Good 0.29 (0.41) Moderate None 0.13 Good Excellent 2 1.20 (1.70) N/A N/A 0.23 (0.33) N/A N/A 0.21 Good N/A3 0.85 (1.20) N/A N/A 0.25 (0.35) N/A N/A 0.20 Good N/A4 1.77 (2.50) Good Poor 0.41 (0.58) Moderate None 0.16 Excellent Excellent Existing 1.77 (2.50) Excellent Excellent 0.46 (0.65) Light Light 0.56 Good GoodBrush Unit Brush Material #1 Brush Material #2 Brush Material #3 Brush Material #4 Existing SN 0042 1.8 (2.5) 1.2 (1.7) 0.85 (1.2) 1.8 (2.5) 1.6 (2.2) EM 0001 2.4 (3.4) N/A N/A 2.3 (3.3) 1.9 (2.7)Motor Torque (N-cm / in-oz) NASA/CP-2010-216272
368 Static Load Distribution under Combined Radial and Thrust Load in Ball Bearings 

It is possible to consider how th e bearing load is distributed among the rolling elements having defined in  other works analytical expressions for geometry of bearings and for contact stress and deformations for a  given ball or roller-raceway contact (point or line loading) in terms of load. In this section, a specific load  distribution consisting of a combined radial and thrust load must be applied to the inner ring of a statically loaded ball bearing so that no tilt is allowed between inner and outer rings.  Let a ball bearing with a number of balls, Z, symmetrically distributed about a pitch circle according to  Figure 1, to be subjected to a combined radial and thrust load, so that a relative axial displacement , δ a,  and a relative radial displacement , δr, between the inner and outer ring raceways may be expected. Let ψ  = 0 to be the angular position of the maximum loaded ball.  Figure 2 shows the initial and final curvat ure centers positions at angular position  ψ, before and after  loading, considering the centers of curvature of the raceway grooves fixed with respect to the corresponding raceway. If δ a and δ r are known, the contact angle at angular position ψ, after the  combined load has been applied, is given by   ()1cos coscosfr nA Aβδψβψδ−+ ⎛⎞=⎜⎟+⎝⎠,      ( 1 )   where A  is the distance between raceway groove cu rvature centers for the unloaded bearing, βf is the  free-contact angle, and δn is the total normal deflection at the contacts.   Also, from Figure 2,   ()f n a A A ββδδ sin sin−+ = ,       ( 2 )   and we can arrive in the expression for the extent of the loading zone , that is given by 

⎪⎭⎪⎬⎫ ⎪⎩⎪⎨⎧ ⎥⎥ ⎦⎤ ⎢⎢ ⎣⎡ −⎟⎟ ⎠⎞ ⎜⎜ ⎝⎛ ⎟⎟ ⎠⎞ ⎜⎜ ⎝⎛+=− − ff a rlAA Aββ δ δψ cossinsin cos cos1 1.   (3)  From (1), the total normal approach between two raceways at angular position ψ, after the combined load  has been applied, can be written as   ()βψ δ ββψ δcoscos1coscosr f n A +⎟⎟ ⎠⎞ ⎜⎜ ⎝⎛− = .      ( 4 )   From Figure 2 and (4) it can be determined that s, the distance between the c enters of the curvature of  the inner and outer ring raceway groov es at any rolling element position ψ, is given by   ()βψ δ ββδ ψcoscos coscosr f nA A s+ = + = .      ( 5 )   From (2) and (5) yields, for ψ = ψj,   ()0cossincos tan =−− − jf j j j r aAβββψ β δ δ ,    j = 1,…, Z.   (6)     NASA/CP-2010-216272
369 From the load-deflection relationship for ball bearings and (4) yields, for ψ = ψj,   2 / 3 coscos1coscos ⎥⎥ ⎦⎤ ⎢⎢ ⎣⎡ +⎟⎟ ⎠⎞ ⎜⎜ ⎝⎛ − = jj r jf nj j A K Qβψ δ ββ ,    j = 1,…, Z.   (7) 

Figure 1.  Ball angular positions in the radial plane that is perpendicular to the bearing’s  axis of rotation, Δψ = 2 π/Z, ψ j = 2π/Z(j−1) 

Figure 2.  Initial and final curvature centers positions at angular position ψ,   with and without applied load 

   Outer raceway groove  curvature center  fixed

βf Initial position,  inner raceway  groove  curvature center   AFinal position,  inner raceway  groove curvature  center  δa   s = A + δn β Acosβf δrcosψ   NASA/CP-2010-216272
370 If a thrust load, Fa, and a radial load, Fr, are applied then, for static equilibrium to exist   ∑ ==Z jj j a Q F 1sinβ,        ( 8 )    jZ jj j r Q F ψ βcos cos 1∑ == .       ( 9 )   Additionally, each of the normal ball load components produces a moment about of  the inner ring center  of mass in the plane that passes through the bearing rotation axis and contains the external radial load  (moments about the other two perpend icular planes are self -equilibrating). For static equilibrium, the  thrust load, F a, and/or the radial load, Fr, must exert a moment, M, about the inner ring center of mass  that must be equal to the sum of the moment s of each rolling element load, that is,   ()[]∑ =− + − =Z jr j j r i j j R Q M 1cos cos sin δ ψ ψ δ β,   (10)  where  ()f i e i D f d Rβcos 5 . 0 2 /−+=   expresses the locus of the centers of the inner ring raceway groove curvature radii. 

Substitution of (7) into (8) yields   0coscos1coscossin 12 / 3 =⎟⎟ ⎠⎞ ⎜⎜ ⎝⎛ +⎟⎟ ⎠⎞ ⎜⎜ ⎝⎛ − −∑ =Z j jj r jf j nj a A K Fβψ δ βββ    ( 1 1 )   Similarly,   0coscos1coscoscos cos 12 / 3 =⎟⎟ ⎠⎞ ⎜⎜ ⎝⎛ +⎟⎟ ⎠⎞ ⎜⎜ ⎝⎛ − −∑ =Z j jj r jf j j nj r A K Fβψ δ βββ ψ    ( 1 2 )   Equations (6), (11) and (12) are Z + 2 simultaneous nonlinear equations with unknowns δa, δr, and βj, j =  1,…, Z. Since Knj are functions of final contact angle, βj, the equations must be solv ed iteratively to yield an  exact solution for  δa, δr and βj. 

Numerical results 

A numerical method (the Newton-Rhapson method) was chosen to solve the simultaneous nonlinear  equations (6), (11) and (12). To show an application of the theory developed in this work, a numerical example is presented. I have chosen the 218 angular-contact ball bearing that was also used by [5]. Thus, the results generated here can be compared to a certain degree with the Harris results.    Figures 3 - 5 show some parameters as functions of the applied thrust load under a radial load of 17,800  N. We can observe a substantial difference between re sults found in this work and those found by Harris,  for a thrust load of 17,800 N.  Figure 3 shows the normal ball loads, Q. Harris found the following ball load magnitudes: 6571; 5765; 3670; and 1200 N, for the balls located at angular positions: |ψ| = 0; 22.5 o; 45o; and 67.5o, respectively,  and found zero ball load magnitudes for the balls located at angular positions | ψ| ≥ 90o (p. 262). This work  found the following ball loads magnitudes: 5997; 5395; 3807; 1820; and 239 N, for the balls located at  angular positions: | ψ| = 0; 22.5o; 45o; 67.5o; and 90o, respectively, and found zero ball load magnitudes for  the balls located at angular positions | ψ| > 90o.  NASA/CP-2010-216272
371 This means that Harris calculation has overestimate d (or underestimated) the normal ball loads for balls  located at angular positions satisfying | ψ| < 45º (| ψ| ≥ 45º); representing an error of 9.56% in the  determination of maximum normal ball load and errors  of 6.86%, –3.59%, –34.06% and –100% in the  determination of normal ball load for the eight balls immediately about the most heavily loaded ball,  respectively.  

Figure 3.  Normal ball load, Q, for 17,800 N radial load, as a function of the thrust load, Fa. 

Figure 4 shows the contact angle, β. While Harris assumed a contact angle magnitude of 40o for all balls  (p. 260), contact angles ranging from 38.2o to 42.7o were found in this work, while ψ were varied from ψ =  0º to ±180o, respectively. This represents errors between 4.71% and –6.28% in the contact angles  determination, meaning that Harris assumption has overestimated (underestimated) the contact angles for balls located at angular positions satisfying | ψ| < 90º (|ψ| ≥ 90º). 

Figure 4.  Contact angle, β, for 17,800 N radial load, as a function of the thrust load, Fa. 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 x 1040100020003000400050006000700080009000 Thrust load,  Fa  [N]Normal Applied Load,    Q  [N]Radial and Thrust Load - 218 Angular-contact Ball Bearing   Fr = 17,800 N = ± 45° = ± 67.5°= ± 90°ψ = 0 = ± 22.5° Harris ψ = 0;      6,571 N = ± 22.5°; 5,765 N = ± 45°; 3,670 N = ± 67.5°; 1,200 N = ± 90°; 0 N 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 x 10451015202530354045 Thrust load, Fa  [N]Contact Angle,   β  [Deg]Radial and Thrust Load - 218 Angular-contact Ball Bearing    ψ = 0  = ± 22.5°  = ± 45°  = ± 67.5°βf  = ± 90° = ± 112.5°  = ± 135°  = ± 157.5°  = 180°Fr = 17,800 N NASA/CP-2010-216272
372 Figure 5 shows the loading zone, ψ l. While Harris found a loading zone of 84.84º (p. 262), this work found  a loading zone of 97.74º. This represents an error of  –13.2% in the loading angle,  meaning that the Harris  calculation has underestimated the effect of the loading. 

Figure 5.  Contact zone, ψl, for 17,800 N radial load, as a function of the thrust load, Fa.     Conclusion   A numerical procedure for computing the internal loading distribution in statically loaded, single-row, angular-contact ball bearings when subjected to a known combined radial and thrust load has been  presented. The combined radial and thrust load must be applied in order to avoid tilting between inner and outer rings. The procedure requi res the iterative solution of Z + 2 simultaneous nonlinear equations  with unknowns δ a, δr, and β j, j = 1,…, Z. Numerical results for a 218 angular-contact ball bearing have  been compared with those from the literature and show  significant differences in the magnitudes of the  ball loads, contact angles, and ex tent of the loading zone. 

References 

1. Stribeck, R. “Ball Bear ings for Various Loads.” Trans. ASME , 29 (1907), 420-463.  2. Sjoväll, H. “The Load Distribution within Ball an d Roller Bearings under Given External Radial and  Axial Load.” Teknisk Tidskrift, Mekanik, häfte 9 (1933), 97-102. (http ://runeberg.org/tektid/1933m/009  9.html)  3. Jones, A. Analysis of Stresses and Deflections . Bristol: New Departure Engineering Data, ©1946.   4. Rumbarger, J. “Thrust Bear ings with Eccentric Loads.” Machine Design (15 February 1962).  5.   Harris, T. Rolling Bearing Analysis . 4th Ed., New York: John Wiley & Sons, Inc., ©2001.  6.  Ricci, M. C., “Ball bearings subjec ted to a variable eccentric thrust load.” DINCON’09 Proceedings of  the 8th Brazilian Conference on Dynamics, Control and Applications , (May 2009). ISBN: 978-8586883-45-3.  7. Ricci, M. C. “Internal loading distribution in statically loaded ball bearings subjected to a combined  radial, thrust, and moment load.” Proceedings of the 60th International Astronautical Congress ,  (October 2009). ISSN 1995-6258. 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 x 104405060708090100110120 Thrust load, Fa  [N]Loading zone,   ψl  [Deg]Radial and Thrust Load - 218 Angular-contact Ball Bearing ψ = 45°= 67.5°= 112.5° = 90°Fr = 17,800 N Harris 84.84° NASA/CP-2010-216272