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+    "viz_data": "{\"columns\":[\"id\",\"x\",\"y\",\"document\",\"document_cleaned\",\"size\",\"category\"],\"index\":[0,1,2,3,4,5,6,7,8,9,10,11,12,13,14,15,16,17,18,19,20,21,22,23,24,25,26,27,28,29,30,31,32,33,34,35,36,37,38,39,40,41,42,43,44,45,46,47,48,49],\"data\":[[\"009bc76a-bd43-11ee-801f-bae7cd9d315f\",5.689907074,6.0673141479,\"181 Developmental Testing of Electric Thr ust Vector Control Systems for   Manned Launch Vehicle Applications \\n\\nLisa B. Bates* and David T. Young** \\n\\nAbstract \\n\\nThis paper describes recent developmental testing to verify the integration of a developmental  electromechanical actuator (EMA) with high rate lithium ion batteries and a cross platform extensible controller. Testing was performed at the Thrust Vector Control Research, Development and Qualification Laboratory at the NASA George C. Marshall Space Flight Center. Electric Thrust Vector Control (ETVC) systems like the EMA may significantly reduce recurring launch costs and complexity compared to heritage systems. Electric actuator mechanisms and control requirements across dissimilar platforms are also discussed with a focus on the similarities leveraged and differences overcome by the cross platform extensible common controller architecture.   Introduction{'source': 'AMS_2012.pdf', 'page': 195}\",\"181 Developmental Testing of Electric Thr ust Vector Control Systems for<br>Manned Launch Vehicle Applications   Lisa B. Bates* and David T. Young**<br>Abstract   This paper describes recent developmental testing to verify the<br>integration of a developmental  electromechanical actuator (EMA) with high rate<br>lithium ion batteries and a cross platform extensible controller. Testing was<br>performed at the Thrust Vector Control Research, Development and Qualification<br>Laboratory at the NASA George C. Marshall Space Flight Center. Electric Thrust<br>Vector Control (ETVC) systems like the EMA may significantly reduce recurring<br>launch costs and complexity compared to heritage systems. Electric actuator<br>mechanisms and control requirements across dissimilar platforms are also<br>discussed with a focus on the similarities leveraged and differences overcome by<br>the cross platform extensible common controller architecture.<br>Introduction{'source': 'AMS_2012.pdf', 'page': 195}\",3,\"Chunks\"],[\"0149266c-bd43-11ee-801f-bae7cd9d315f\",3.2121088505,7.6814665794,\"Figure 15.  Stroke Testing of the final guide  design under high loading conditions (load is  equivalent to 3.2-g radial and 5.4 times the  maximum flight moment){'source': 'AMS_2012.pdf', 'page': 265}\",\"Figure 15.  Stroke Testing of the final guide  design under high loading<br>conditions (load is  equivalent to 3.2-g radial and 5.4 times the  maximum<br>flight moment){'source': 'AMS_2012.pdf', 'page': 265}\",3,\"Chunks\"],[\"03ae7b3c-bd43-11ee-801f-bae7cd9d315f\",7.4775462151,8.7771959305,\"from pushing the the green Drive Pin out of the Toroidal Bushing. When the ERM releases the  central rotor, the drive spring is free to expand and pushes the Drive Pin out of the Toroidal  Bushing and into a guide hole in the anchor fitting. To reset the actuator, the the Drive Pin is  manually pulled back through the Toroidal Bushing (v ia the Reset Linkage a nd the Rotor Arm) and  connected to the reset ERM.  The titanium Anchor FIttings and Toroidal Bushings had a Tiodize\\u00ae type II (Teflon\\u00ae impregnated) finish. The Drive Pins are made of aluminum bronze, CDA 63020 per AMS 4590B. A light film of Braycote{'source': 'AMS_2012.pdf', 'page': 408}\",\"from pushing the the green Drive Pin out of the Toroidal Bushing. When the ERM<br>releases the  central rotor, the drive spring is free to expand and pushes the<br>Drive Pin out of the Toroidal  Bushing and into a guide hole in the anchor<br>fitting. To reset the actuator, the the Drive Pin is  manually pulled back<br>through the Toroidal Bushing (v ia the Reset Linkage a nd the Rotor Arm) and<br>connected to the reset ERM.  The titanium Anchor FIttings and Toroidal Bushings<br>had a Tiodize\\u00ae type II (Teflon\\u00ae impregnated) finish. The Drive Pins are made of<br>aluminum bronze, CDA 63020 per AMS 4590B. A light film of Braycote{'source':<br>'AMS_2012.pdf', 'page': 408}\",3,\"Chunks\"],[\"05928484-bd43-11ee-801f-bae7cd9d315f\",4.4773106575,10.9182357788,\"the first  seque nce, were perform ed a nd th e  mech anism was pla ced into vacuum cyclin g. Tabl e 3 shows number of op eration s at st eps d uring the  second lifetime simulatio n seque nce.   Table 3.  HCM Life-T est Second Seq uence \\n\\nOper ation Seque nce 2 Ops  Break-in CW Spin 56,200  Pre-Lifetest Fu nctional Testing 138,180  Vibration Test 138,180  Thermal F unctional Testing 704,340  Software Ve rification 727,570  Vacuum Ope ration 3,573,250  Post-Lifetest Fun ctional Testing 3,691,832  Additional Repea tability  Testing 4,276,882                    \\n\\nFigure 13a & 13b.  Proto-Qualifica tion HCM Conta mination:  L ocking Fea ture on Mech anism  Scre w (left) and an Alum inum and Ch romate Par ticle Remov ed from the Bearing (righ t)  23{'source': 'AMS_2004.pdf', 'page': 37}\",\"the first  seque nce, were perform ed a nd th e  mech anism was pla ced into<br>vacuum cyclin g. Tabl e 3 shows number of op eration s at st eps d uring the<br>second lifetime simulatio n seque nce.   Table 3.  HCM Life-T est Second Seq<br>uence   Oper ation Seque nce 2 Ops  Break-in CW Spin 56,200  Pre-Lifetest Fu<br>nctional Testing 138,180  Vibration Test 138,180  Thermal F unctional Testing<br>704,340  Software Ve rification 727,570  Vacuum Ope ration 3,573,250  Post-<br>Lifetest Fun ctional Testing 3,691,832  Additional Repea tability  Testing<br>4,276,882                      Figure 13a & 13b.  Proto-Qualifica tion HCM Conta<br>mination:  L ocking Fea ture on Mech anism  Scre w (left) and an Alum inum and<br>Ch romate Par ticle Remov ed from the Bearing (righ t)  23{'source':<br>'AMS_2004.pdf', 'page': 37}\",3,\"Chunks\"],[\"06e5a7c6-bd43-11ee-801f-bae7cd9d315f\",6.7834515572,9.8055076599,\"high fo rces a pplied to  the l ocking pins by the latc h mechanism. De sign changes to the pl ate were ma de to   increa se stiffness at the l ocking pi n mounting p oints, and thi s chang e elimin ated plate b endin g as a  probl em. \\n\\nLatch Point E ngag ement   Becau se of manufa cturing tolerances and th e curved m otion of  the spring beams,  some dimen sional  allowan ce must be mad e at the beam-t o-latch pin co ntact point. In the initial desig n, the allowa nce wa s  gene rous, and the holding ability of the l atch in the Z di rection was augmente d by frictional forces. In fact,   mating surfaces were tex tured by grit blastin g in order to  enhance fri ction. It was fou nd in  testing  that  texturing a ctually aggravat ed gallin g of the mating surface s, and th at desig n feat ure was delet ed. Smooth  surfa ces were use d inste ad, with su rface treatment for hardening. Tolera nces were tightene d, and it wa s  then po ssible to redu ce th e dimen sion of the pocket  in the spri ng beam in whi ch the lo cking pin se ats.  The upp er and lowe r sho ulders of the  pocket offer pos itive re straint of the locking pin which is n ot  depe ndent o n friction.  105{'source': 'AMS_2004.pdf', 'page': 119}\",\"high fo rces a pplied to  the l ocking pins by the latc h mechanism. De sign<br>changes to the pl ate were ma de to   increa se stiffness at the l ocking pi n<br>mounting p oints, and thi s chang e elimin ated plate b endin g as a  probl em.<br>Latch Point E ngag ement   Becau se of manufa cturing tolerances and th e curved<br>m otion of  the spring beams,  some dimen sional  allowan ce must be mad e at<br>the beam-t o-latch pin co ntact point. In the initial desig n, the allowa nce wa<br>s  gene rous, and the holding ability of the l atch in the Z di rection was<br>augmente d by frictional forces. In fact,   mating surfaces were tex tured by<br>grit blastin g in order to  enhance fri ction. It was fou nd in  testing  that<br>texturing a ctually aggravat ed gallin g of the mating surface s, and th at<br>desig n feat ure was delet ed. Smooth  surfa ces were use d inste ad, with su<br>rface treatment for hardening. Tolera nces were tightene d, and it wa s  then po<br>ssible to redu ce th e dimen sion of the pocket  in the spri ng beam in whi ch<br>the lo cking pin se ats.  The upp er and lowe r sho ulders of the  pocket offer<br>pos itive re straint of the locking pin which is n ot  depe ndent o n friction.<br>105{'source': 'AMS_2004.pdf', 'page': 119}\",3,\"Chunks\"],[\"06e5a8f2-bd43-11ee-801f-bae7cd9d315f\",6.1723618507,10.4475269318,\"Coordinate 3, Cran kshaft L oads (Fig. 4)   The inte rmediate sh aft driv es a cran kshaft with an gular Coordinate \\u03b83, having  two b earing drag to rques  T31B and T 32B. These two drag torques could be referenced directly to the drive shaft, skipping the  intermediate shaft ( \\u03b82), by their di splacement ratio t o the drive  shaft (\\u03b83\\/\\u03b81). However,  this task is more  system aticall y organized by first collecting the  loads from th e crank and pisto n, and th en referencing their  sum from the crankshaft ( \\u03b83) to the driveshaft ( \\u03b81).  \\n\\nCoordinates 4, 5, and 6, Linka ge Bea ring Load s and Piston Loa ds (Fig. 4)   Although th e friction to rque from th e inte rmedi ate crank bearing T4B also acts directly on the crankshaft,  this be aring\\u2019s displ acemen t, \\u03a64, is relative to its two mating crank linkages. This relative displacement, \\u03a64,  is gre ater that  the displa cement \\u03b83 of the lower link relative to the cran kshaft, in the crank position shown.   \\n\\n\\u03a64 = \\u03b83 + \\u03b85       (7) \\n\\nThe fri ction t orque of this cra nk bearing acts on both the upp er crank and th e lower crank. Thu s, the{'source': 'AMS_2004.pdf', 'page': 127}\",\"Coordinate 3, Cran kshaft L oads (Fig. 4)   The inte rmediate sh aft driv es a<br>cran kshaft with an gular Coordinate \\u03b83, having  two b earing drag to rques<br>T31B and T 32B. These two drag torques could be referenced directly to the drive<br>shaft, skipping the  intermediate shaft ( \\u03b82), by their di splacement ratio t o<br>the drive  shaft (\\u03b83\\/\\u03b81). However,  this task is more  system aticall y<br>organized by first collecting the  loads from th e crank and pisto n, and th en<br>referencing their  sum from the crankshaft ( \\u03b83) to the driveshaft ( \\u03b81).<br>Coordinates 4, 5, and 6, Linka ge Bea ring Load s and Piston Loa ds (Fig. 4)<br>Although th e friction to rque from th e inte rmedi ate crank bearing T4B also<br>acts directly on the crankshaft,  this be aring\\u2019s displ acemen t, \\u03a64, is<br>relative to its two mating crank linkages. This relative displacement, \\u03a64,  is<br>gre ater that  the displa cement \\u03b83 of the lower link relative to the cran<br>kshaft, in the crank position shown.     \\u03a64 = \\u03b83 + \\u03b85       (7)   The fri ction<br>t orque of this cra nk bearing acts on both the upp er crank and th e lower<br>crank. Thu s, the{'source': 'AMS_2004.pdf', 'page': 127}\",3,\"Chunks\"],[\"0784b726-bd43-11ee-801f-bae7cd9d315f\",6.8542079926,9.7067556381,\"\\u2022 Latch 5: Dim ensions: 102 mm (4\\u201d)L X 1 02mm (4 \\u201d)W X 191mm (7. 5\\u201d)H             Mass = 1.4 kg   Material s for Major Comp onents    \\u2022 ACME Sc rew: CRES 17-4PH  \\u2022 Barden 107H Ball B earings: CRES 4 40C  \\u2022 Latch Housing: Al 7075-T7 3  \\n\\n133{'source': 'AMS_2004.pdf', 'page': 147}\",\"\\u2022 Latch 5: Dim ensions: 102 mm (4\\u201d)L X 1 02mm (4 \\u201d)W X 191mm (7. 5\\u201d)H<br>Mass = 1.4 kg   Material s for Major Comp onents    \\u2022 ACME Sc rew: CRES 17-4PH<br>\\u2022 Barden 107H Ball B earings: CRES 4 40C  \\u2022 Latch Housing: Al 7075-T7 3<br>133{'source': 'AMS_2004.pdf', 'page': 147}\",3,\"Chunks\"],[\"08323004-bd43-11ee-801f-bae7cd9d315f\",4.3582348824,10.8563108444,\"material s. Fi nally, the a bility to weld  titanium a llowed the suspen sion stru ctural  comp onents to b e  optimize d for stren gth and flexibility. Eight of the ten suspension tube mem bers we re welde d. The de sire  to increa se the Ti-6AL-4V  from the anneale d to a solution treate d and aged (STA) state was resisted.  While th e ST A pro cess would incre ase the strength of the titanium from 90 0 MPa (13 0 ksi) t o 1100 MPa  (160 ksi), the  weld seam s would rem ain in the anne aled co ndition, creating a n obvious a nd unacce ptable  weak li nk that  could only be mitigated if the STA pr ocess was performe d after welded. T he possibility that  the weld me mbers would distort signi ficantly durin g the STA pro cess due to th eir thin walle d constructio n  was deem ed too risky to accept. Therefore, all the su spension tub e membe rs were kept in their ann ealed  condition. \\n\\nStructu ral Me mber F abrication  The de sire to cre ate a su spen sion that efficiently ab sorbs e nergy leads to st ructural m embers th at are  thin wall ed box beam s. A b ox beam design is a m ass efficient ge ometry for compone nts subjected to  both   bendi ng and torsio nal loads. The  beams are also tapered wherever po ssible t o incre ase m ass saving s.{'source': 'AMS_2004.pdf', 'page': 203}\",\"material s. Fi nally, the a bility to weld  titanium a llowed the suspen sion<br>stru ctural  comp onents to b e  optimize d for stren gth and flexibility. Eight<br>of the ten suspension tube mem bers we re welde d. The de sire  to increa se the<br>Ti-6AL-4V  from the anneale d to a solution treate d and aged (STA) state was<br>resisted.  While th e ST A pro cess would incre ase the strength of the titanium<br>from 90 0 MPa (13 0 ksi) t o 1100 MPa  (160 ksi), the  weld seam s would rem ain<br>in the anne aled co ndition, creating a n obvious a nd unacce ptable  weak li nk<br>that  could only be mitigated if the STA pr ocess was performe d after welded. T<br>he possibility that  the weld me mbers would distort signi ficantly durin g the<br>STA pro cess due to th eir thin walle d constructio n  was deem ed too risky to<br>accept. Therefore, all the su spension tub e membe rs were kept in their ann<br>ealed  condition.   Structu ral Me mber F abrication  The de sire to cre ate a<br>su spen sion that efficiently ab sorbs e nergy leads to st ructural m embers th<br>at are  thin wall ed box beam s. A b ox beam design is a m ass efficient ge<br>ometry for compone nts subjected to  both   bendi ng and torsio nal loads. The<br>beams are also tapered wherever po ssible t o incre ase m ass saving<br>s.{'source': 'AMS_2004.pdf', 'page': 203}\",3,\"Chunks\"],[\"0a6ab382-bd43-11ee-801f-bae7cd9d315f\",4.5527009964,11.0192298889,\"The de sign of  the butterfly element s that engage a nd become prelo aded ag ainst a rock was a formida ble  task. It is  necessary to ac tually pr eload the ground structure of the RAT ag ainst the rock, effectively{'source': 'AMS_2004.pdf', 'page': 293}\",\"The de sign of  the butterfly element s that engage a nd become prelo aded ag<br>ainst a rock was a formida ble  task. It is  necessary to ac tually pr eload the<br>ground structure of the RAT ag ainst the rock, effectively{'source':<br>'AMS_2004.pdf', 'page': 293}\",3,\"Chunks\"],[\"0dd5be86-bd43-11ee-801f-bae7cd9d315f\",7.563829422,8.7669439316,\"which the corer would be mounted by reducing mass and required electrical connections.   Other design considerations were addressed prior to finalizing the overall concept of the integrated corer  design. A decision was made to use a single motor to drive a spring-loaded rotary-percussive cam mechanism, much like the original LSAS drill. Although it was desirable to have the flexibility of independent control over the hammer and rotary functi ons in the manner of the drilling breadboard fixture,  the single-motor rotary-percussive mechanism offered simplicity and low mass. An additional motor was incorporated in the design to accomplish core break-off. While it appeared possible to leverage the rotary-percussive motor to achieve this function and save mass, it was much more straightforward to add a second motor. See Figures 6 and 7 for CAD views of the prototype design.    NASA\\/CP-2010-216272{'source': 'AMS_2010.pdf', 'page': 38}\",\"which the corer would be mounted by reducing mass and required electrical<br>connections.   Other design considerations were addressed prior to finalizing<br>the overall concept of the integrated corer  design. A decision was made to use<br>a single motor to drive a spring-loaded rotary-percussive cam mechanism, much<br>like the original LSAS drill. Although it was desirable to have the flexibility<br>of independent control over the hammer and rotary functi ons in the manner of<br>the drilling breadboard fixture,  the single-motor rotary-percussive mechanism<br>offered simplicity and low mass. An additional motor was incorporated in the<br>design to accomplish core break-off. While it appeared possible to leverage the<br>rotary-percussive motor to achieve this function and save mass, it was much more<br>straightforward to add a second motor. See Figures 6 and 7 for CAD views of the<br>prototype design.    NASA\\/CP-2010-216272{'source': 'AMS_2010.pdf', 'page': 38}\",3,\"Chunks\"],[\"0f8d8e8e-bd43-11ee-801f-bae7cd9d315f\",5.6056170464,6.1294298172,\"174 Acknowledgements \\n\\nThe authors and co-authors of this paper express their gratitude to the failure investigation teams. The  collaborative efforts of NASA, Boeing, NESC, Battelle, The Aerospace Corporation, L-3 S&N, and independent consultants resulted in comprehensive inve stigations that resulted in identifying the most  probable cause for both anomalie s. Accolades we re expressed by Boeing and NASA upper management  many times for the team work, leadership and efficiency of the teams identified below:  \\n\\n   CMG1 Failure Investigation Team     NASA\\/CP-2010-216272{'source': 'AMS_2010.pdf', 'page': 190}\",\"174 Acknowledgements   The authors and co-authors of this paper express their<br>gratitude to the failure investigation teams. The  collaborative efforts of<br>NASA, Boeing, NESC, Battelle, The Aerospace Corporation, L-3 S&N, and<br>independent consultants resulted in comprehensive inve stigations that resulted<br>in identifying the most  probable cause for both anomalie s. Accolades we re<br>expressed by Boeing and NASA upper management  many times for the team work,<br>leadership and efficiency of the teams identified below:       CMG1 Failure<br>Investigation Team     NASA\\/CP-2010-216272{'source': 'AMS_2010.pdf', 'page':<br>190}\",3,\"Chunks\"],[\"14a7ada0-bd43-11ee-801f-bae7cd9d315f\",7.1994800568,7.1037535667,\"(sync) cable provides redundancy for the dampers and deployment springs, allowing for system  deployment in the event of a spring or damper failure. The sync cable also helps maintain the ratio of the two hinge angles to approximately 2:1, as the elbow and shoulder hinges open 180 and 90 degrees respectively during deployment.  HGAS contains five mechanisms:  three Launch Restraint Mechanisms (LRM\\u2019s), a Lower Boom Assembly (LBA), and the gimbal assembly. The upper boom connects the LBA to the gimbal assembly as shown in  Figure 1. The entire HGAS assembly and associated mechanism are supported on an all-aluminum honeycomb plate. The LRMs and hinge line designs were developed from heritage Solar Dynamics Observatory hardware. Launch Restraint Mechanism The HGAS uses three LRMs  (known as the A, B, and C devices) to restrain the upper  boom and the gimbal assembly to the mounting plate prior to  deployment. Each LRM is composed of a latch rod, securing  the upper boom or gimbal  assembly to two spring-loaded jaws. Each jaw is attached to a  non-explosive actuator (NEA).  After firing, the NEAs release the latch rods allowing the system to deploy. As designed, LRM C  releases first, followed four seconds later by the LRMs A and B simultaneously. Deployment commences once all LRMs are released. Kick-off springs at LRM B and LRM C assist in separating the HGAS assembly from the LRMs.  Lower Boom Assembly The LBA, shown in Figure 2, is made up of the two deployment hinges; the elbow hinge and shoulder{'source': 'AMS_2014.pdf', 'page': 62}\",\"(sync) cable provides redundancy for the dampers and deployment springs,<br>allowing for system  deployment in the event of a spring or damper failure. The<br>sync cable also helps maintain the ratio of the two hinge angles to<br>approximately 2:1, as the elbow and shoulder hinges open 180 and 90 degrees<br>respectively during deployment.  HGAS contains five mechanisms:  three Launch<br>Restraint Mechanisms (LRM\\u2019s), a Lower Boom Assembly (LBA), and the gimbal<br>assembly. The upper boom connects the LBA to the gimbal assembly as shown in<br>Figure 1. The entire HGAS assembly and associated mechanism are supported on an<br>all-aluminum honeycomb plate. The LRMs and hinge line designs were developed<br>from heritage Solar Dynamics Observatory hardware. Launch Restraint Mechanism<br>The HGAS uses three LRMs  (known as the A, B, and C devices) to restrain the<br>upper  boom and the gimbal assembly to the mounting plate prior to  deployment.<br>Each LRM is composed of a latch rod, securing  the upper boom or gimbal<br>assembly to two spring-loaded jaws. Each jaw is attached to a  non-explosive<br>actuator (NEA).  After firing, the NEAs release the latch rods allowing the<br>system to deploy. As designed, LRM C  releases first, followed four seconds<br>later by the LRMs A and B simultaneously. Deployment commences once all LRMs are<br>released. Kick-off springs at LRM B and LRM C assist in separating the HGAS<br>assembly from the LRMs.  Lower Boom Assembly The LBA, shown in Figure 2, is made<br>up of the two deployment hinges; the elbow hinge and shoulder{'source':<br>'AMS_2014.pdf', 'page': 62}\",3,\"Chunks\"],[\"157a4896-bd43-11ee-801f-bae7cd9d315f\",6.2788877487,6.0096197128,\"As it could not be actually measured, the stow bias was the least certain value in the ADAMS model. To  finalize calibration of the model, the stow bias was iterated until the latch time of both hinges closely matched the latching (i.e. fully deployed) times in the most current g-negated tests. The hinges would not latch at the exact same time due to compliance in the sync cable, allow\\/g76\\/g81\\/g74\\/g3\\/g87\\/g75\\/g72\\/g3\\/g86\\/g80\\/g68\\/g79\\/g79\\/g3\\/g507\\/g3\\/g68\\/g81\\/g74\\/g79\\/g72\\/g3between the two hinges.  The final calibration to the  ADAMS model was done  using the post T-Vac, gnegated, deployment data.  The hinge angles of the post  T-vac test and the  corresponding calibrated ADAMS model predictions are compared in Figure 6. The 3 N (0.7 lbf) of stow bias used to calibrate the model compares favorably to the 2 to 4.4 N (0.5 to 1.0 lbf) stow bias estimated by the test engineers prior to the test. This calibrated model became the baseline ADAMS model. This baseline model, run as a \\u201con-orbit\\u201d (i.e. zero gravity  and g-negation system  model elements disabled) would be used to predict onorbit performance of the flight configuration HGAS.  Potential Deployment Interference After the HGAS subsystem was integrated to the spacecraft, a potential deployment interference issue{'source': 'AMS_2014.pdf', 'page': 68}\",\"As it could not be actually measured, the stow bias was the least certain value<br>in the ADAMS model. To  finalize calibration of the model, the stow bias was<br>iterated until the latch time of both hinges closely matched the latching (i.e.<br>fully deployed) times in the most current g-negated tests. The hinges would not<br>latch at the exact same time due to compliance in the sync cable, allow\\/g76\\/g81\\/<br>g74\\/g3\\/g87\\/g75\\/g72\\/g3\\/g86\\/g80\\/g68\\/g79\\/g79\\/g3\\/g507\\/g3\\/g68\\/g81\\/g74\\/g79\\/g72\\/g3betwe<br>en the two hinges.  The final calibration to the  ADAMS model was done  using<br>the post T-Vac, gnegated, deployment data.  The hinge angles of the post  T-vac<br>test and the  corresponding calibrated ADAMS model predictions are compared in<br>Figure 6. The 3 N (0.7 lbf) of stow bias used to calibrate the model compares<br>favorably to the 2 to 4.4 N (0.5 to 1.0 lbf) stow bias estimated by the test<br>engineers prior to the test. This calibrated model became the baseline ADAMS<br>model. This baseline model, run as a \\u201con-orbit\\u201d (i.e. zero gravity  and<br>g-negation system  model elements disabled) would be used to predict onorbit<br>performance of the flight configuration HGAS.  Potential Deployment Interference<br>After the HGAS subsystem was integrated to the spacecraft, a potential<br>deployment interference issue{'source': 'AMS_2014.pdf', 'page': 68}\",3,\"Chunks\"],[\"16394cfa-bd43-11ee-801f-bae7cd9d315f\",3.0079112053,8.5311737061,\"Finally, we show that some of the lubricant that is displaced between the ball and race during run-in operation can be recovered during rest, and we measure the rate of recovery for one example. 1.0 Introduction Bearing life and performance is critically dependent on lubricant. Heat transfer is also dependent on  lubricant in space, therefore the two are linked. This paper will show bearing thermal properties depend on lubricant and its quantity, then, show how the conductance measurements can be used to infer lubricant behavior. The requirements for operation of space mechanisms present bearings a very different thermal  environment than mechanisms used in a terrestrial environment. In terrestrial applications, convection dominates the cooling mechanism. If air is not enough to cool it, the bearing is typically flooded with  lubricant for additional cooling. Thus, bearing thermal conductance tends to be a second or third order effect in most terrestrial applications. However, in the vacuum of space, essentially no air is present and flooding with lubricant is not feasible. Furthermore, in most cases, the bearing must operate with parched lubricant quantities and perform for  years under these conditions. In the absence of convection, bearing raceway temperatures are a productof bearing thermal conductance, heat generation, and the operational environmental temperature. In most  *The Aerospace Corporation, El Segundo, CA **The Aerospace Corporation, Colorado Springs, CO{'source': 'AMS_2014.pdf', 'page': 129}\",\"Finally, we show that some of the lubricant that is displaced between the ball<br>and race during run-in operation can be recovered during rest, and we measure<br>the rate of recovery for one example. 1.0 Introduction Bearing life and<br>performance is critically dependent on lubricant. Heat transfer is also<br>dependent on  lubricant in space, therefore the two are linked. This paper will<br>show bearing thermal properties depend on lubricant and its quantity, then, show<br>how the conductance measurements can be used to infer lubricant behavior. The<br>requirements for operation of space mechanisms present bearings a very different<br>thermal  environment than mechanisms used in a terrestrial environment. In<br>terrestrial applications, convection dominates the cooling mechanism. If air is<br>not enough to cool it, the bearing is typically flooded with  lubricant for<br>additional cooling. Thus, bearing thermal conductance tends to be a second or<br>third order effect in most terrestrial applications. However, in the vacuum of<br>space, essentially no air is present and flooding with lubricant is not<br>feasible. Furthermore, in most cases, the bearing must operate with parched<br>lubricant quantities and perform for  years under these conditions. In the<br>absence of convection, bearing raceway temperatures are a productof bearing<br>thermal conductance, heat generation, and the operational environmental<br>temperature. In most  *The Aerospace Corporation, El Segundo, CA **The Aerospace<br>Corporation, Colorado Springs, CO{'source': 'AMS_2014.pdf', 'page': 129}\",3,\"Chunks\"],[\"17aef0c6-bd43-11ee-801f-bae7cd9d315f\",6.9785842896,9.4924850464,\"014 HBR 179 Ratchet 0.89 40 4.02 189509 6535 C 015** HBR 179 Ratchet 0.78 40 4.56 222700 7679 F 016 HBR 179 Ratchet 0.81 41 4.42 208367 7185 A 017 HBR 179 Ratchet 0.90 39 3.85 188187 6489 C 018 JPL 179 Ratchet 1.15 37 3.26 149220 5146 A 019 JPL 179 Ratchet 1.01 39 3.44 167460 5774 D 020 JPL 179 Ratchet 0.88 39 3.78 198538 6846 D 021 JPL 179 Ratchet 1.04 37 3.48 173618 5987 A Lessons Learned A number of useful lessons were learned from the design and testing of the RANCOR drill. As with any  design, there is still room for improvement, but in the end the drill was more than capable of performing  coring tasks in medium to low strength rock targets. Also as requirements for Mars Sample Return (MSR)mature, there may be more mass and volume available to the drill design that can be utilized to increase the reliability and robustness. Lessons learned from the RANCOR drill include the following: 1. In this case, the cost and simplicity of an off-the-shelf sprag clutch versus the design and build of a  custom ratchet and pawl system led to the decision to use the sprag clutch. Although there was nothing functionally wrong with the sprag in this design, it enabled a degree of freedom that should have been locked out during the release of the hammer on the RANCOR. Therefore, the sprag mechanism was replaced with a ratchet and pawl system. The result was a large improvement in core quality (from D through F grades to A and B grades).{'source': 'AMS_2014.pdf', 'page': 244}\",\"014 HBR 179 Ratchet 0.89 40 4.02 189509 6535 C 015** HBR 179 Ratchet 0.78 40<br>4.56 222700 7679 F 016 HBR 179 Ratchet 0.81 41 4.42 208367 7185 A 017 HBR 179<br>Ratchet 0.90 39 3.85 188187 6489 C 018 JPL 179 Ratchet 1.15 37 3.26 149220 5146<br>A 019 JPL 179 Ratchet 1.01 39 3.44 167460 5774 D 020 JPL 179 Ratchet 0.88 39<br>3.78 198538 6846 D 021 JPL 179 Ratchet 1.04 37 3.48 173618 5987 A Lessons<br>Learned A number of useful lessons were learned from the design and testing of<br>the RANCOR drill. As with any  design, there is still room for improvement, but<br>in the end the drill was more than capable of performing  coring tasks in medium<br>to low strength rock targets. Also as requirements for Mars Sample Return<br>(MSR)mature, there may be more mass and volume available to the drill design<br>that can be utilized to increase the reliability and robustness. Lessons learned<br>from the RANCOR drill include the following: 1. In this case, the cost and<br>simplicity of an off-the-shelf sprag clutch versus the design and build of a<br>custom ratchet and pawl system led to the decision to use the sprag clutch.<br>Although there was nothing functionally wrong with the sprag in this design, it<br>enabled a degree of freedom that should have been locked out during the release<br>of the hammer on the RANCOR. Therefore, the sprag mechanism was replaced with a<br>ratchet and pawl system. The result was a large improvement in core quality<br>(from D through F grades to A and B grades).{'source': 'AMS_2014.pdf', 'page':<br>244}\",3,\"Chunks\"],[\"18fa2f4a-bd43-11ee-801f-bae7cd9d315f\",3.5367951393,6.9706912041,\"plate. This configuration was chosen instead of the fully clamped one to simulate the actual balancing configuration in which the dynamometer\\u2019s own mode is in the vicinity of 30 Hz versus the 120 Hz+ in the fully clamped configuration. Three measurements were taken. In the first measurement, the data from the load cells is channeled only to an independent secondary data acquisition system (VXI system). In the second test, the data from the load cells is channeled to the BDMS data acquisition (processor to be used  for balancing) and simultaneously to the VXI through a set of \\u201cT\\u201d connections. The third test has both data  acquisition systems in place except that the data acquisition card of the BDMS was turned off. A digital filter of 10 Hz was applied to the data sets. An example of dynamometer noise floor measurement for the  Y-axis is shown in Figure 5. The maximum noise of 0.00175 lb (0.00778 N )is measured and is found to  be less than the 0.004 lb (0.018 N) needed to carry out accurate measurements to meet GMI spin  balance requirements. Harmonic stinger test This test is intended to calibrate the measurements performed by the dynamometer as a system. The  dynamometer is excited at a frequency of 10 Hz using an MTB 50-lb (220-N) stinger. A calibrated load cell is placed at the interface between the stinger and the dynamometer to measure the input force. The  310{'source': 'AMS_2014.pdf', 'page': 326}\",\"plate. This configuration was chosen instead of the fully clamped one to<br>simulate the actual balancing configuration in which the dynamometer\\u2019s own mode<br>is in the vicinity of 30 Hz versus the 120 Hz+ in the fully clamped<br>configuration. Three measurements were taken. In the first measurement, the data<br>from the load cells is channeled only to an independent secondary data<br>acquisition system (VXI system). In the second test, the data from the load<br>cells is channeled to the BDMS data acquisition (processor to be used  for<br>balancing) and simultaneously to the VXI through a set of \\u201cT\\u201d connections. The<br>third test has both data  acquisition systems in place except that the data<br>acquisition card of the BDMS was turned off. A digital filter of 10 Hz was<br>applied to the data sets. An example of dynamometer noise floor measurement for<br>the  Y-axis is shown in Figure 5. The maximum noise of 0.00175 lb (0.00778 N )is<br>measured and is found to  be less than the 0.004 lb (0.018 N) needed to carry<br>out accurate measurements to meet GMI spin  balance requirements. Harmonic<br>stinger test This test is intended to calibrate the measurements performed by<br>the dynamometer as a system. The  dynamometer is excited at a frequency of 10 Hz<br>using an MTB 50-lb (220-N) stinger. A calibrated load cell is placed at the<br>interface between the stinger and the dynamometer to measure the input force.<br>The  310{'source': 'AMS_2014.pdf', 'page': 326}\",3,\"Chunks\"],[\"199b50f0-bd43-11ee-801f-bae7cd9d315f\",6.6779203415,10.0152788162,\"(%)DUTY  CYCLE  (msec)VOLTAGE  (Vdc)TORQUE  PRIMARY  (CW) Nm (lb-in)TORQUE  PRIMARY  (CCW) Nm (lb-in)TORQUE RED.  (CW) Nm (lb-in)TORQUE  RED.  (CCW) Nm (lb-in) 1 4 40 23.13 21.5(190) 21.5(190) 21.5(190) 21.5(190) 1 5.1 51 23.13 21.5(190) 22.6(200) 21.5(190) 22.0(195) 1 8 80 23.13 22.6(200) 22.0(195) 21.5(190) 22.0(195) 1 8 80 27.39 22.6(200) 22.6(200) 22.0(195) 22.0(195) 10 40 40 23.13 21.5(190) 20.9(185) 21.5(190) 20.3(180) 10 51 51 23.13 20.3(180) 19.8(175) 20.9(185) 19.8(175) 10 80 80 23.13 21.5(190) 20.9(185) 21.5(190) 20.9(185) 10 80 80 27.39 21.5(190) 20.9(185) 22.0(195) 21.5(190) 358{'source': 'AMS_2014.pdf', 'page': 374}\",\"(%)DUTY  CYCLE  (msec)VOLTAGE  (Vdc)TORQUE  PRIMARY  (CW) Nm (lb-in)TORQUE<br>PRIMARY  (CCW) Nm (lb-in)TORQUE RED.  (CW) Nm (lb-in)TORQUE  RED.  (CCW) Nm (lb-<br>in) 1 4 40 23.13 21.5(190) 21.5(190) 21.5(190) 21.5(190) 1 5.1 51 23.13<br>21.5(190) 22.6(200) 21.5(190) 22.0(195) 1 8 80 23.13 22.6(200) 22.0(195)<br>21.5(190) 22.0(195) 1 8 80 27.39 22.6(200) 22.6(200) 22.0(195) 22.0(195) 10 40<br>40 23.13 21.5(190) 20.9(185) 21.5(190) 20.3(180) 10 51 51 23.13 20.3(180)<br>19.8(175) 20.9(185) 19.8(175) 10 80 80 23.13 21.5(190) 20.9(185) 21.5(190)<br>20.9(185) 10 80 80 27.39 21.5(190) 20.9(185) 22.0(195) 21.5(190) 358{'source':<br>'AMS_2014.pdf', 'page': 374}\",3,\"Chunks\"],[\"1a56d1e0-bd43-11ee-801f-bae7cd9d315f\",3.2640552521,7.8682575226,\"actual magnitude of the pull-in torque at a desired pulse rate is more of an iterative process. A relatively  simple method that is not as laborious of continually increasing the torque test by test is to run the actuator at the desired pulse rate and increase the torque until the unit pulls-out of synchronous  operation. While the unit is \\u201cbuzzing\\u201d in this pulled out condition, decrease the brake torque until the actuator regains synchronous operation. Stop the unit and allow the actuator to return to room  temperature, then test the pull-in torque at the torque value that returned the actuator to synchronous operation. A final minor adjustment may be necessary, but this value should be a negligible difference to  the actual pull-in torque value.  Now that the performance at room temperature has been characterized, it may be desired to simulate or  test the performance at maximum temperature. It is actually simple to simulate high temperature  performance at ambient room temperature by calculating the motor resistance and power input as described in Equation 7 and Appendix A. With the reduced power calculated, simply adjust the power input to the supply to apply the high temperature input power, and duplicate the dynamic tests described above.  Adjusting the supply voltage cannot simulate testing an actuator at colder temperatures. While the  electromagnetically generated torques are proportional in decreasing temperature, increased lube viscosity could dominate at colder temperatures and increase torque losses greater than the increased generated torque through reduced resistance. Testing inside a temperature chamber may be the only  alternative, but if the cold temperature values are well within a lubricant\\u2019s rating, room temperature  performance may be acceptable.    Conclusion Linear interpretation to simulate stepper motor performance when introduced to load inertia, given motor{'source': 'AMS_2014.pdf', 'page': 403}\",\"actual magnitude of the pull-in torque at a desired pulse rate is more of an<br>iterative process. A relatively  simple method that is not as laborious of<br>continually increasing the torque test by test is to run the actuator at the<br>desired pulse rate and increase the torque until the unit pulls-out of<br>synchronous  operation. While the unit is \\u201cbuzzing\\u201d in this pulled out<br>condition, decrease the brake torque until the actuator regains synchronous<br>operation. Stop the unit and allow the actuator to return to room  temperature,<br>then test the pull-in torque at the torque value that returned the actuator to<br>synchronous operation. A final minor adjustment may be necessary, but this value<br>should be a negligible difference to  the actual pull-in torque value.  Now that<br>the performance at room temperature has been characterized, it may be desired to<br>simulate or  test the performance at maximum temperature. It is actually simple<br>to simulate high temperature  performance at ambient room temperature by<br>calculating the motor resistance and power input as described in Equation 7 and<br>Appendix A. With the reduced power calculated, simply adjust the power input to<br>the supply to apply the high temperature input power, and duplicate the dynamic<br>tests described above.  Adjusting the supply voltage cannot simulate testing an<br>actuator at colder temperatures. While the  electromagnetically generated<br>torques are proportional in decreasing temperature, increased lube viscosity<br>could dominate at colder temperatures and increase torque losses greater than<br>the increased generated torque through reduced resistance. Testing inside a<br>temperature chamber may be the only  alternative, but if the cold temperature<br>values are well within a lubricant\\u2019s rating, room temperature  performance may<br>be acceptable.    Conclusion Linear interpretation to simulate stepper motor<br>performance when introduced to load inertia, given motor{'source':<br>'AMS_2014.pdf', 'page': 403}\",3,\"Chunks\"],[\"1a56d2e4-bd43-11ee-801f-bae7cd9d315f\",7.0366444588,9.5628852844,\"Figure 9.  Fishbone Diagram Failure Scenario Supporting Elements The following key findings related to the root cause investigation elements were utilized to devise a most  probable failure scenario. 1.Test Video and Strain Gauge Data : Immediately after application of the qualification levels, the SM  side assembly (aft cap, actuator housing, and forward cap) rotated clockwise (looking SM to umbilical  side). Lagging this rotation, the secondary piston was observed to rotate counterclockwise. Inspection of data from strain gauges mounted on the secondary piston and visual indicators of mounted accelerometers indicated a rotation of approximately 90 degrees. Strain gauge data with indicators of tooling hole orientation relative to the Y-axis applied load can be found in Figure 10. Approximately 27  seconds after application of qualification levels (242 seconds) there was a noticeable decrease in noise and strut dynamic response. No noticeable damage was observed, so the test was continued to the full duration at 395 seconds. 2.Forward Lug Locking Patch Design : Load requirements for the locking patch were not defined prior  to developmental testing. Designers did not anticipate any applied loosening torque to aid in the patch  sizing in part due to the fact that the strut would be in compression during vibration testing.Additionally, it is not standard practice to perform supporting analyses of locking patch capability.Locking patches are not intended to serve as torque reaction features; rather, they should be used to reduce the rate of preload loss in a joint. The prevailing torque requirement was less than the minimum recommended running torques specified for fine thread series threads 38.1 mm (1.500 in) in diameter or less. The locking patch vendor indicated the patch size was small relative to the thread  size and pitch to which they were applied.   3.Joint Characteristics : The threads utilized for testing had an as-designed preload limited to 25{'source': 'AMS_2014.pdf', 'page': 414}\",\"Figure 9.  Fishbone Diagram Failure Scenario Supporting Elements The following<br>key findings related to the root cause investigation elements were utilized to<br>devise a most  probable failure scenario. 1.Test Video and Strain Gauge Data :<br>Immediately after application of the qualification levels, the SM  side assembly<br>(aft cap, actuator housing, and forward cap) rotated clockwise (looking SM to<br>umbilical  side). Lagging this rotation, the secondary piston was observed to<br>rotate counterclockwise. Inspection of data from strain gauges mounted on the<br>secondary piston and visual indicators of mounted accelerometers indicated a<br>rotation of approximately 90 degrees. Strain gauge data with indicators of<br>tooling hole orientation relative to the Y-axis applied load can be found in<br>Figure 10. Approximately 27  seconds after application of qualification levels<br>(242 seconds) there was a noticeable decrease in noise and strut dynamic<br>response. No noticeable damage was observed, so the test was continued to the<br>full duration at 395 seconds. 2.Forward Lug Locking Patch Design : Load<br>requirements for the locking patch were not defined prior  to developmental<br>testing. Designers did not anticipate any applied loosening torque to aid in the<br>patch  sizing in part due to the fact that the strut would be in compression<br>during vibration testing.Additionally, it is not standard practice to perform<br>supporting analyses of locking patch capability.Locking patches are not intended<br>to serve as torque reaction features; rather, they should be used to reduce the<br>rate of preload loss in a joint. The prevailing torque requirement was less than<br>the minimum recommended running torques specified for fine thread series threads<br>38.1 mm (1.500 in) in diameter or less. The locking patch vendor indicated the<br>patch size was small relative to the thread  size and pitch to which they were<br>applied.   3.Joint Characteristics : The threads utilized for testing had an as-<br>designed preload limited to 25{'source': 'AMS_2014.pdf', 'page': 414}\",3,\"Chunks\"],[\"1e4324b6-bd43-11ee-801f-bae7cd9d315f\",6.2449645996,5.9957976341,\"workenvelope constraints, Thisvolume wasroughly 7cm(2.75in)longitudinal by10cm(4in)lateralby 3.2cm(1.25in)tall.Evenwithinthoselimits,itwasstrongly desired tokeepactuator volume toa minimum, sincefutureapplications mayrequire multiple actuators ontheglove.Somecomponents ofthe system (e.g.powersupply, microprocessor, etc.)couldbelocated remotely inthebackpack (PLSS). Safetyisaprimeconsideration inspacesuit systems. Thisconcern ledtotherequirement ofno penetrations ofthepressure bladder forthissystem; allcomponents mustbeexternal. Thissystem must alsoavoidcreating ahazardous temperature insidetheglove.Amaximum temperature riseatthebase oftheactuator (backofglove)of40\\u00b0C(72\\u00b0F)overambient (poweroff)wassetasalimit.Thesystem mustbedesigned tofailsafe.Specifically, nocredible failuremodecancompromise theintegrity ofthe pressure suitorprevent theoperator fromperforming themanual tasksneeded toingress theairlock. Nostrictlimitonpowerrequirements wasinitiallyset.Aneventual flightsystem mustprovide sufficient powerforasix-hour EVAduration, andbeabletorejecttheheatdissipated internally. Thedesignintent atthisphasewastominimize powerconsumption giventherequired performance andotherconstraints. 9O{'source': 'AMS_2000.pdf', 'page': 104}\",\"workenvelope constraints, Thisvolume wasroughly 7cm(2.75in)longitudinal<br>by10cm(4in)lateralby 3.2cm(1.25in)tall.Evenwithinthoselimits,itwasstrongly<br>desired tokeepactuator volume toa minimum, sincefutureapplications mayrequire<br>multiple actuators ontheglove.Somecomponents ofthe system (e.g.powersupply,<br>microprocessor, etc.)couldbelocated remotely inthebackpack (PLSS).<br>Safetyisaprimeconsideration inspacesuit systems. Thisconcern ledtotherequirement<br>ofno penetrations ofthepressure bladder forthissystem; allcomponents<br>mustbeexternal. Thissystem must alsoavoidcreating ahazardous temperature<br>insidetheglove.Amaximum temperature riseatthebase oftheactuator<br>(backofglove)of40\\u00b0C(72\\u00b0F)overambient (poweroff)wassetasalimit.Thesystem<br>mustbedesigned tofailsafe.Specifically, nocredible failuremodecancompromise<br>theintegrity ofthe pressure suitorprevent theoperator fromperforming themanual<br>tasksneeded toingress theairlock. Nostrictlimitonpowerrequirements<br>wasinitiallyset.Aneventual flightsystem mustprovide sufficient powerforasix-hour<br>EVAduration, andbeabletorejecttheheatdissipated internally. Thedesignintent<br>atthisphasewastominimize powerconsumption giventherequired performance<br>andotherconstraints. 9O{'source': 'AMS_2000.pdf', 'page': 104}\",3,\"Chunks\"],[\"2037c79a-bd43-11ee-801f-bae7cd9d315f\",2.7576487064,8.6010456085,\"Figure 3 Materials: Asdiscussed earlier, allbearing components werefabricated fromnon-magnetic materials. Considerable effortsweremadetoobtainiron-free variants ofthemany\\\"common\\\" materials employed intheconstruction ofthegimbal. Lubrication: Nolubricants areemployed. Duetothesemi-cryogenic operational temperatures, no wetlubricants weresuitable. Theuseofjeweled bearings precluded theneedforanyadditional lubrication. Thecombination ofthematerials employed inthebearing resultsinacoefficient of frictionof0.15. Vibration: Thespringwithinthebearing isapotential singlepointfailuresothatattention mustbe concentrated onthisimportant aspectofthedesign. Areliability prediction wasperformed toverifya verylowfailurerateintheapplication ofthespring. Thespringuses0.2-mm diameter wirewithameancoildiameter of1.32mmdiameter. Thereare6 activecoilswithapitchof0.33mm.Thematerial isBeryllium-copper. Thespringrateforacompression springis: R=(modulus ofrigidity)(wire diameter_ 8(meanspringdiameter)3(Active coils) R=4.24Ibs.\\/in Thespringconcentration factoriswherer=coildiameter towirediameter 193{'source': 'AMS_2000.pdf', 'page': 207}\",\"Figure 3 Materials: Asdiscussed earlier, allbearing components werefabricated<br>fromnon-magnetic materials. Considerable effortsweremadetoobtainiron-free<br>variants ofthemany\\\"common\\\" materials employed intheconstruction ofthegimbal.<br>Lubrication: Nolubricants areemployed. Duetothesemi-cryogenic operational<br>temperatures, no wetlubricants weresuitable. Theuseofjeweled bearings precluded<br>theneedforanyadditional lubrication. Thecombination ofthematerials employed<br>inthebearing resultsinacoefficient of frictionof0.15. Vibration:<br>Thespringwithinthebearing isapotential singlepointfailuresothatattention mustbe<br>concentrated onthisimportant aspectofthedesign. Areliability prediction<br>wasperformed toverifya verylowfailurerateintheapplication ofthespring.<br>Thespringuses0.2-mm diameter wirewithameancoildiameter of1.32mmdiameter.<br>Thereare6 activecoilswithapitchof0.33mm.Thematerial isBeryllium-copper.<br>Thespringrateforacompression springis: R=(modulus ofrigidity)(wire diameter_<br>8(meanspringdiameter)3(Active coils) R=4.24Ibs.\\/in Thespringconcentration<br>factoriswherer=coildiameter towirediameter 193{'source': 'AMS_2000.pdf', 'page':<br>207}\",3,\"Chunks\"],[\"2037c952-bd43-11ee-801f-bae7cd9d315f\",7.6906294823,7.6343193054,\"_verArm Flexures TorqueTube StopArmon DriveHub Figure11aand11b.TheCoverDriveMechanism opensandcloses thecanister cover.  ArmsAttached OI _toCanister Cover ..._._'I- (Torque TubeNotShown' Brackets Attached Canister Base (Support PlatesNotShown) GearmotorO Figure 12.0 b= 0 Bearing O Preload Spring_ r_ TheCoverDriveMechanism cross-sectionCable Spool 211{'source': 'AMS_2000.pdf', 'page': 225}\",\"_verArm Flexures TorqueTube StopArmon DriveHub<br>Figure11aand11b.TheCoverDriveMechanism opensandcloses thecanister cover.<br>ArmsAttached OI _toCanister Cover ..._._'I- (Torque TubeNotShown' Brackets<br>Attached Canister Base (Support PlatesNotShown) GearmotorO Figure 12.0 b= 0<br>Bearing O Preload Spring_ r_ TheCoverDriveMechanism cross-sectionCable Spool<br>211{'source': 'AMS_2000.pdf', 'page': 225}\",3,\"Chunks\"],[\"22d00684-bd43-11ee-801f-bae7cd9d315f\",4.8223829269,10.9381322861,\":':::'_ii':':i_:i_i_.:_:i_2! 1:2_2121;2_ii++222!2:2_!22;i! Figure7aFreonTFprocessed bearing scan ........................... .++......... ....;............. .........,..,..+.+,.,.. .................. ,........................................ +........... o.................... ....+,.:,., ......... ._,.,...... :......... ;....,+.++,,._+.,. \\/ ....... ,...... ,......................... ;........... .....,+ .............................................................. ,...................,.........,........................... _.2--\\\"'--.u-_%J Figure7c.HFE-7100 processed bearing scan........ ,,..:........ +........... \\u2022...:,..,: ............ ,,+. _k \\/ .)Jlill'_ 111 ___L\\\"!....i....i...._....i....i...._....i....:....i...._....:, Figure7b.VertrelXFprocessed bearingscan Alltracesat50rpm.Allscansrepresent slightlygreaterthanonebearinginner shaftrevolution. Figure7.Interimtorquetracesforpreloaded bearing pairsafterapproximately 2.8.10srevolutions !+i..........................i........._.........i........._.........::........._........_........_........--.--:.........i...........\\u2022..!....!...._....:....:....!....:....!....i....;....!....i* 1\\\"6.4oz'inl \\\"\\\"_...._....i....i........i....!....i\\\"'+!....i....i.....\\\" Figure8b.VertrelXFprocessed bearingscan{'source': 'AMS_2001.pdf', 'page': 41}\",\":':::'_ii':':i_:i_i_.:_:i_2! 1:2_2121;2_ii++222!2:2_!22;i!<br>Figure7aFreonTFprocessed bearing scan ........................... .++.........<br>....;............. .........,..,..+.+,.,.. ..................<br>,........................................ +........... o....................<br>....+,.:,., ......... ._,.,...... :......... ;....,+.++,,._+.,. \\/ .......<br>,...... ,......................... ;........... .....,+<br>..............................................................<br>,...................,.........,........................... _.2--\\\"'--.u-_%J<br>Figure7c.HFE-7100 processed bearing scan........ ,,..:........ +...........<br>\\u2022...:,..,: ............ ,,+. _k \\/ .)Jlill'_ 111<br>___L\\\"!....i....i...._....i....i...._....i....:....i...._....:,<br>Figure7b.VertrelXFprocessed bearingscan Alltracesat50rpm.Allscansrepresent<br>slightlygreaterthanonebearinginner shaftrevolution.<br>Figure7.Interimtorquetracesforpreloaded bearing pairsafterapproximately<br>2.8.10srevolutions !+i..........................i........._.........i........._.<br>........::........._........_........_........--.--<br>:.........i...........\\u2022..!....!...._....:....:....!....:....!....i....;....!....<br>i* 1\\\"6.4oz'inl \\\"\\\"_...._....i....i........i....!....i\\\"'+!....i....i.....\\\"<br>Figure8b.VertrelXFprocessed bearingscan{'source': 'AMS_2001.pdf', 'page': 41}\",3,\"Chunks\"],[\"245a24ee-bd43-11ee-801f-bae7cd9d315f\",6.9780030251,9.4912605286,\"between thebackstop ofthedoublehingeandthetapered wedgeofthelatchpin.Thetapered latchpin willnotallowanyplay.Itwilldothisbyspringing axiallyforward totakeupanygapthatwouldbeopened byplayinthehingeassembly. Thisgapisusuallycaused byclearances duetotolerances. TestsandResults Theobjectofthistestwastoseewhether thetapered latchpinassembly trulyhasremoved alloftheplay fromthemechanism. Table1showstheplayattheendofatypicalzeroplayhingelatchforside1and side2withthehingelatchassembly locked intheopenposition. Figure3shows thesetupofthe measuring equipment. Playwasmeasured intherotational direction. Thisisthedirection thatthehinge wouldnormally rotate.Theplayintheopened, rotational direction isaverylowvalue0.00254 -0.00381 mm(0.0001-0.00015 in)butisnotzero.Thiswascaused bythefactthatintestingitacertainamount of forcewasapplied tothehinge,about30g(1oz),sothatsomereading wouldappear onthedial indicator. Itwasnecessary todothistomakesuretheweightofthehingewasnotpreventing thehinge frommoving freely.Whentheforcewasapplied, itprobably pushed thelatchpinbackward inanaxial direction, causing somelooseness. Thereading wouldbeclosertozeroiflessforcewereused.Afterthe playinthehingewasmeasured, eachsideofthehingewastakenapartandtheaxlesandholeswere{'source': 'AMS_2001.pdf', 'page': 126}\",\"between thebackstop ofthedoublehingeandthetapered wedgeofthelatchpin.Thetapered<br>latchpin willnotallowanyplay.Itwilldothisbyspringing axiallyforward<br>totakeupanygapthatwouldbeopened byplayinthehingeassembly. Thisgapisusuallycaused<br>byclearances duetotolerances. TestsandResults Theobjectofthistestwastoseewhether<br>thetapered latchpinassembly trulyhasremoved alloftheplay fromthemechanism.<br>Table1showstheplayattheendofatypicalzeroplayhingelatchforside1and<br>side2withthehingelatchassembly locked intheopenposition. Figure3shows<br>thesetupofthe measuring equipment. Playwasmeasured intherotational direction.<br>Thisisthedirection thatthehinge wouldnormally rotate.Theplayintheopened,<br>rotational direction isaverylowvalue0.00254 -0.00381 mm(0.0001-0.00015<br>in)butisnotzero.Thiswascaused bythefactthatintestingitacertainamount of<br>forcewasapplied tothehinge,about30g(1oz),sothatsomereading wouldappear onthedial<br>indicator. Itwasnecessary todothistomakesuretheweightofthehingewasnotpreventing<br>thehinge frommoving freely.Whentheforcewasapplied, itprobably pushed<br>thelatchpinbackward inanaxial direction, causing somelooseness. Thereading<br>wouldbeclosertozeroiflessforcewereused.Afterthe playinthehingewasmeasured,<br>eachsideofthehingewastakenapartandtheaxlesandholeswere{'source': 'AMS_2001.pdf',<br>'page': 126}\",3,\"Chunks\"],[\"282473ea-bd43-11ee-801f-bae7cd9d315f\",4.0847411156,6.1787433624,\"70 pitch scale, located at a radius of 31. 06 mm from  the mirror rotation axis. This sensor has a 1. 2-nm  resolution,  which is thus equivalent to 38. 6 nrad of mirror  rotation. The encoder was originally meant for  beam acquisition purposes, but can also be used as a feedback  sensor (as an alternative to the external  interferometer).  \\n\\nFigure 3.  Picture of the realized two -stepper IFPM. (Photo: TNO \\/ Gert Witvoet) \\n\\nThe piezosteppers are fed by four high- voltage space- qualified  analog amplifiers, one for each of the four   phases of the actuators. The voltage waveforms are generated by a dSpace data acquisition system with  a 16- bit D\\/A converter; t he encoder (via a 24- bit digital encoder  interface) and the interferometer (via a  16-bit A\\/D converter)  are connected to the same dSpace system. This  system offers a rapid prototyping  environment in MATLAB \\/Simulink, which provides great flexibility in meas urement  possibilities and  controller design.  \\n\\nSystem Behavior  \\n\\nThe motion of the legs of the walking actuator is determined by the voltage distribution along the four  phases  as a function of time. Although the open- loop motion  will never be perfectly linear, the exact  shape of these voltage waveforms has a large influence on the velocity  variatio ns during an actuator  cycle [6]. For example, the horizontal and vertical motion of the first set of legs can be approximated by{'source': 'AMS_2016.pdf', 'page': 84}\",\"70 pitch scale, located at a radius of 31. 06 mm from  the mirror rotation axis.<br>This sensor has a 1. 2-nm  resolution,  which is thus equivalent to 38. 6 nrad<br>of mirror  rotation. The encoder was originally meant for  beam acquisition<br>purposes, but can also be used as a feedback  sensor (as an alternative to the<br>external  interferometer).    Figure 3.  Picture of the realized two -stepper<br>IFPM. (Photo: TNO \\/ Gert Witvoet)   The piezosteppers are fed by four high-<br>voltage space- qualified  analog amplifiers, one for each of the four   phases<br>of the actuators. The voltage waveforms are generated by a dSpace data<br>acquisition system with  a 16- bit D\\/A converter; t he encoder (via a 24- bit<br>digital encoder  interface) and the interferometer (via a  16-bit A\\/D converter)<br>are connected to the same dSpace system. This  system offers a rapid prototyping<br>environment in MATLAB \\/Simulink, which provides great flexibility in meas<br>urement  possibilities and  controller design.    System Behavior    The motion<br>of the legs of the walking actuator is determined by the voltage distribution<br>along the four  phases  as a function of time. Although the open- loop motion<br>will never be perfectly linear, the exact  shape of these voltage waveforms has<br>a large influence on the velocity  variatio ns during an actuator  cycle [6].<br>For example, the horizontal and vertical motion of the first set of legs can be<br>approximated by{'source': 'AMS_2016.pdf', 'page': 84}\",3,\"Chunks\"],[\"2bdc3824-bd43-11ee-801f-bae7cd9d315f\",5.1426267624,5.8477778435,\"199 A New Architecture for Absolute  Optical Encoders  \\n\\nTimothy  Malcolm*, John Beasley * and Mike Jumper * \\n\\nAbstract    BEI Precision Systems & Space has developed an encoder technology, nanoSeries, that can calibrate itself in-situ and correct most of the common causes of error in typical encoders. The new nanoSeries ARA  design has detailed health and status readouts that can definitively indicate when a re- calibration is in order .  The re- calibration process can be done on- orbit if desired. The units can accommodate either full  revolutions or limited angle sweeps, and the principles are also applicable to linear encoders.  \\n\\nIntroduction  \\n\\nOptical  encoders  manufactured  by BEI Precision   Systems & Space  (BEI)  have  been  used  in space since  the  earliest  days  of space flight.  The combination of  low we ight and high resolution  relative to electromagnetic   resolvers  made them  an obvious  choice  for many  applications.  Optical encoders have typically been of  only a  few types.  \\n\\nThe primary  type of optical encoder  selected for commercial  and industrial  applications  has been the   incremental  encoder.  This type of encoder  requires  a return to an index  or home pulse to index a counter,   which  then counts  the number  of \\u2018incremental\\u2019  pulses  or bits that pass  by. This is a very simple  and robust   concept  but it has some disadvantages  for space,  primarily  that if power  should go off or the counter  upset s,{'source': 'AMS_2016.pdf', 'page': 213}\",\"199 A New Architecture for Absolute  Optical Encoders    Timothy  Malcolm*, John<br>Beasley * and Mike Jumper *   Abstract    BEI Precision Systems & Space has<br>developed an encoder technology, nanoSeries, that can calibrate itself in-situ<br>and correct most of the common causes of error in typical encoders. The new<br>nanoSeries ARA  design has detailed health and status readouts that can<br>definitively indicate when a re- calibration is in order .  The re- calibration<br>process can be done on- orbit if desired. The units can accommodate either full<br>revolutions or limited angle sweeps, and the principles are also applicable to<br>linear encoders.    Introduction    Optical  encoders  manufactured  by BEI<br>Precision   Systems & Space  (BEI)  have  been  used  in space since  the<br>earliest  days  of space flight.  The combination of  low we ight and high<br>resolution  relative to electromagnetic   resolvers  made them  an obvious<br>choice  for many  applications.  Optical encoders have typically been of  only a<br>few types.    The primary  type of optical encoder  selected for commercial  and<br>industrial  applications  has been the   incremental  encoder.  This type of<br>encoder  requires  a return to an index  or home pulse to index a counter,<br>which  then counts  the number  of \\u2018incremental\\u2019  pulses  or bits that pass  by.<br>This is a very simple  and robust   concept  but it has some disadvantages  for<br>space,  primarily  that if power  should go off or the counter  upset<br>s,{'source': 'AMS_2016.pdf', 'page': 213}\",3,\"Chunks\"],[\"2c94b944-bd43-11ee-801f-bae7cd9d315f\",7.6605796814,7.5263586044,\"Slip Ring  Elevation Rotary Joint   Base  bracket   HRM   NEA  MLI Antenna Inertia Dummy   Elevation Stage    Azimuth Stage   Azimuth   Stepper Motor  Waveguide   Elevation Rotary Joint{'source': 'AMS_2016.pdf', 'page': 262}\",\"Slip Ring  Elevation Rotary Joint   Base  bracket   HRM   NEA  MLI Antenna<br>Inertia Dummy   Elevation Stage    Azimuth Stage   Azimuth   Stepper Motor<br>Waveguide   Elevation Rotary Joint{'source': 'AMS_2016.pdf', 'page': 262}\",3,\"Chunks\"],[\"377b9ea4-bd43-11ee-801f-bae7cd9d315f\",6.9092364311,6.7066822052,\"tosurvive launch. Mounting thegimbaltotheopticalbenchalsoservedtosimplify thedesignbyreducing thenumber ofparts.Italsosimplified instrument assembly andintegration sincethedevelopment ofthe opticalbenchassembly wasdecoupled fromthedevelopment ofthegimbalassembly andbothcouldbe thenconnected ordisconnected whenrequired withminimal effort. NASA\\/C P--2002-211506 224{'source': 'AMS_2002.pdf', 'page': 240}\",\"tosurvive launch. Mounting thegimbaltotheopticalbenchalsoservedtosimplify<br>thedesignbyreducing thenumber ofparts.Italsosimplified instrument assembly<br>andintegration sincethedevelopment ofthe opticalbenchassembly wasdecoupled<br>fromthedevelopment ofthegimbalassembly andbothcouldbe thenconnected<br>ordisconnected whenrequired withminimal effort. NASA\\/C P--2002-211506<br>224{'source': 'AMS_2002.pdf', 'page': 240}\",3,\"Chunks\"],[\"3b365804-bd43-11ee-801f-bae7cd9d315f\",3.8486495018,10.7294206619,\"Mechanical Engineering Department, directed byProf.HariDharan. Theprepreg layerswerelaidbyhandonapolished aluminum mandrel, thenwrapped withafilmof PTFETeflon. Instead ofusinganautoclave orvacuum bagging, alengthofthick-walled neoprene heatshrinktubewasplacedaroundtheuncured tube.Theshrink temperature ofthetubingisthesameasthecuretemperature oftheepoxyinthe prepreg (175\\u00b0C). Theentireassembly wasbakedfortwohoursandcooled. Theshrink 79{'source': 'AMS_1998.pdf', 'page': 91}\",\"Mechanical Engineering Department, directed byProf.HariDharan. Theprepreg<br>layerswerelaidbyhandonapolished aluminum mandrel, thenwrapped withafilmof<br>PTFETeflon. Instead ofusinganautoclave orvacuum bagging, alengthofthick-walled<br>neoprene heatshrinktubewasplacedaroundtheuncured tube.Theshrink temperature<br>ofthetubingisthesameasthecuretemperature oftheepoxyinthe prepreg (175\\u00b0C).<br>Theentireassembly wasbakedfortwohoursandcooled. Theshrink 79{'source':<br>'AMS_1998.pdf', 'page': 91}\",3,\"Chunks\"],[\"3c5bc5f2-bd43-11ee-801f-bae7cd9d315f\",4.1136646271,10.7853841782,\"environment iscrucial. Standard graphite brushes thatoperate acceptably atsealevel provide verypoorperformance inlowpressure andvacuum conditions. Thisis because ofthelackofhumidity andatmosphere thatsupports thedevelopment ofan oxidelayeronthecommutator, thisfilmreduces boththemechanical andelectrical wearofthebrush. Another contributor toelectrical arcingistheinductive energystoredinthecoil.This formofelectrical arcingwascontrolled byincreasing thenumber ofcommutator barsas highaspossible. Themaximum number ofcommutator barsfeasible intheapplication waseleven. Thehighernumber ofcommutator barsthelessinductance andhence lessinductive energy. Thisreduces theelectrical stressonthebrush. Brusharcingwasthenfurtherreduced bytheaddition ofaceramic capacitor inparallel withthebrushassemblies. Theceramic capacitor actstosuppress arcingofthe brushes byreducing theeffective sourceimpedance athighfrequencies. 145{'source': 'AMS_1998.pdf', 'page': 157}\",\"environment iscrucial. Standard graphite brushes thatoperate acceptably<br>atsealevel provide verypoorperformance inlowpressure andvacuum conditions.<br>Thisis because ofthelackofhumidity andatmosphere thatsupports thedevelopment<br>ofan oxidelayeronthecommutator, thisfilmreduces boththemechanical andelectrical<br>wearofthebrush. Another contributor toelectrical arcingistheinductive<br>energystoredinthecoil.This formofelectrical arcingwascontrolled byincreasing<br>thenumber ofcommutator barsas highaspossible. Themaximum number ofcommutator<br>barsfeasible intheapplication waseleven. Thehighernumber ofcommutator<br>barsthelessinductance andhence lessinductive energy. Thisreduces theelectrical<br>stressonthebrush. Brusharcingwasthenfurtherreduced bytheaddition ofaceramic<br>capacitor inparallel withthebrushassemblies. Theceramic capacitor actstosuppress<br>arcingofthe brushes byreducing theeffective sourceimpedance athighfrequencies.<br>145{'source': 'AMS_1998.pdf', 'page': 157}\",3,\"Chunks\"],[\"3d062632-bd43-11ee-801f-bae7cd9d315f\",4.6170172691,10.8186187744,\"reaction wheelincorporating thesehybridbearings revealed nodamage afterthe bearings weresubjected toastresslevelof3780MPa. Introduction Thesuccessful design ofahigh-cycle spacecraft mechanism thatemploys bearings requires thatthebearing materials haveatleasttwoproperties: highrollingcontact fatigue (RCF)resistance tomeetoperational dynamic cycling requirements, and adequate staticloadcapacity tosurvive launch loads. Inrecentyears,testresults havebeenreported forhybridbearings consisting ofSi3N4ballsandsteelraceways in commercial machine toolspindles, inmilitary bearing applications, 1andintheSpace Shuttle mainengine liquidoxygen andfuelturbopumps. 2Suchhybridbearings appear toprovide goodfatigue performance andtoavoidmetal-to-metal contact, which, inturn,retards theonsetoflubricant degradation. Mostofthehybridbearing results reported todateuseexisting bearing steels(52100, 440C,M50,M50Nil), although Cronidur 30steelisplanned foruseintheSpace Shuttle fuelturbopump bearings. 3However, bearing materials areofinterest thatcanoperate underhigher stresswithlongerlife.Recently, apowder-metallurgy highspeedM62toolsteel, calledVIMREX20, orCRU20, hasemerged asagoodcandidate forimproved bearings duetoitshighhardness (HRC66-67) andwearresistance, finecarbide structure, andimproved RCFperformance. Earlyball-on-rod fatigue testshavebeen *TheAerospace Corporation, ElSegundo, CA MPBCorp.,Keene,NH +TimkenCo.,Canton, OH{'source': 'AMS_1998.pdf', 'page': 249}\",\"reaction wheelincorporating thesehybridbearings revealed nodamage afterthe<br>bearings weresubjected toastresslevelof3780MPa. Introduction Thesuccessful<br>design ofahigh-cycle spacecraft mechanism thatemploys bearings requires<br>thatthebearing materials haveatleasttwoproperties: highrollingcontact fatigue<br>(RCF)resistance tomeetoperational dynamic cycling requirements, and adequate<br>staticloadcapacity tosurvive launch loads. Inrecentyears,testresults<br>havebeenreported forhybridbearings consisting ofSi3N4ballsandsteelraceways in<br>commercial machine toolspindles, inmilitary bearing applications, 1andintheSpace<br>Shuttle mainengine liquidoxygen andfuelturbopumps. 2Suchhybridbearings appear<br>toprovide goodfatigue performance andtoavoidmetal-to-metal contact, which,<br>inturn,retards theonsetoflubricant degradation. Mostofthehybridbearing results<br>reported todateuseexisting bearing steels(52100, 440C,M50,M50Nil), although<br>Cronidur 30steelisplanned foruseintheSpace Shuttle fuelturbopump bearings.<br>3However, bearing materials areofinterest thatcanoperate underhigher<br>stresswithlongerlife.Recently, apowder-metallurgy highspeedM62toolsteel,<br>calledVIMREX20, orCRU20, hasemerged asagoodcandidate forimproved bearings<br>duetoitshighhardness (HRC66-67) andwearresistance, finecarbide structure,<br>andimproved RCFperformance. Earlyball-on-rod fatigue testshavebeen *TheAerospace<br>Corporation, ElSegundo, CA MPBCorp.,Keene,NH +TimkenCo.,Canton, OH{'source':<br>'AMS_1998.pdf', 'page': 249}\",3,\"Chunks\"],[\"40d597fc-bd43-11ee-801f-bae7cd9d315f\",6.6094326973,6.2874503136,\"I. im 1.1:'il,Q I I,_.___PANCAM &NAVCAM MASTSTOWED STRONGBACK t-1' \\/ \\/\\/--suspENSioN ROBOTICA_ _=_===\\/=__ _ ,.s_X__'_\\\"7 ASSEMBLY Figure 2.FIDORoverwithMastStowed andInstrument ArmDeployed 125{'source': 'AMS_1999.pdf', 'page': 139}\",\"I. im 1.1:'il,Q I I,_.___PANCAM &NAVCAM MASTSTOWED STRONGBACK t-1' \\/<br>\\/\\/--suspENSioN ROBOTICA_ _=_===\\/=__ _ ,.s_X__'_\\\"7 ASSEMBLY Figure<br>2.FIDORoverwithMastStowed andInstrument ArmDeployed 125{'source':<br>'AMS_1999.pdf', 'page': 139}\",3,\"Chunks\"],[\"44927f90-bd43-11ee-801f-bae7cd9d315f\",6.834312439,6.4797415733,\"TheAir-bearing Deployment Rig General Thedevelopment program, supported bytheDutchgovernment, hasbeenstructured intothefollowing phases: 1.Requirements definition ofnewdeployment rig 2.Breadboardtestprogram (singlebearing) 3.Development trolleytestprogram (onestandard trolley) 4.Replacement ofacomplete trolleyset(4trolleys) Forthedevelopment ofthenewdeployment rigthefollowing setof\\\"toplevel\\\" requirements havebeenderived: 1.Thefunctionality oftheoldrigshallatleastbecovered bythenewdesign, whichmeansthatcriticalinterface dimensions, suchasstowed interpanel spacing (50mm)shallbemaintained. 2.Theexisting rigstructure canbeusedwithonlyminormodifications. 3.Thetrolleydisturbance forceinthetransverse direction oftherigshallbeless than0.01Natatrolleysuspension loadof100N(0.1%). 4.Thetrolleydisturbance forceinthelongitudinal direction oftherigshallbe lessthan0.1Natatrolleysuspension loadof100N(1%). 5.Thetrolleys shallbelightweighttokeepdynamic disturbance forceslow. 6.Thedeployment rigimpactonthecleanroomenvironment shallbe minimized. 7.Themaintenance required bytherigshallbeminimized. 8.Theprocurement costsofatrolleyshallnotexceed thatoftheconventional trolley. Considering theserequirements, thenewdeployment rigtechnology hadtofitintothe existing riginfrastructure (railsystem andpanelspacing instowed wingsituation).{'source': 'AMS_1999.pdf', 'page': 421}\",\"TheAir-bearing Deployment Rig General Thedevelopment program, supported<br>bytheDutchgovernment, hasbeenstructured intothefollowing phases: 1.Requirements<br>definition ofnewdeployment rig 2.Breadboardtestprogram (singlebearing)<br>3.Development trolleytestprogram (onestandard trolley) 4.Replacement ofacomplete<br>trolleyset(4trolleys) Forthedevelopment ofthenewdeployment rigthefollowing<br>setof\\\"toplevel\\\" requirements havebeenderived: 1.Thefunctionality<br>oftheoldrigshallatleastbecovered bythenewdesign, whichmeansthatcriticalinterface<br>dimensions, suchasstowed interpanel spacing (50mm)shallbemaintained.<br>2.Theexisting rigstructure canbeusedwithonlyminormodifications.<br>3.Thetrolleydisturbance forceinthetransverse direction oftherigshallbeless<br>than0.01Natatrolleysuspension loadof100N(0.1%). 4.Thetrolleydisturbance<br>forceinthelongitudinal direction oftherigshallbe<br>lessthan0.1Natatrolleysuspension loadof100N(1%). 5.Thetrolleys<br>shallbelightweighttokeepdynamic disturbance forceslow. 6.Thedeployment<br>rigimpactonthecleanroomenvironment shallbe minimized. 7.Themaintenance required<br>bytherigshallbeminimized. 8.Theprocurement costsofatrolleyshallnotexceed<br>thatoftheconventional trolley. Considering theserequirements, thenewdeployment<br>rigtechnology hadtofitintothe existing riginfrastructure (railsystem<br>andpanelspacing instowed wingsituation).{'source': 'AMS_1999.pdf', 'page': 421}\",3,\"Chunks\"],[\"4688dcea-bd43-11ee-801f-bae7cd9d315f\",6.1103167534,10.4711399078,\"<45de_\\/min. Bending 3.96x104N\\u00b0m\\/deg (2.0x107in\\u00b0lbf\\/rad) Axial1.75x105N\\/cm(1.0x105Ibf\\/in) Combined bending moments abouttwoorthogonal axes2825N\\u00b0m(25,000in\\u00b0tbf), torsional load113N.m(1000in\\u00b0lbf),shearload3145N(707Ibf),axialload2224N (500Ibf) Combined bending moment abouttwoorthogonal axes,1130Nom(10,000in\\u00b0lbf) aboutoneaxis,10,170N.m(90,000in\\u00b0lbf)abouttheotheraxis(thebullgearis lockedfromrotatingusingalockrackwhileloadsareapplied), torsional load5085 N(45T000in\\u00b0lbf)_shearload3145N(707Ibf),axialload2224N(500Ibf) Random vibration_ composite Glevel8.8grms Operating Temperature -40to+60\\u00b0C(-40to+140\\u00b0F) 107kg(235Ib) Storage >10years,operating >10years 103{'source': 'AMS_1997.pdf', 'page': 119}\",\"<45de_\\/min. Bending 3.96x104N\\u00b0m\\/deg (2.0x107in\\u00b0lbf\\/rad)<br>Axial1.75x105N\\/cm(1.0x105Ibf\\/in) Combined bending moments abouttwoorthogonal<br>axes2825N\\u00b0m(25,000in\\u00b0tbf), torsional<br>load113N.m(1000in\\u00b0lbf),shearload3145N(707Ibf),axialload2224N (500Ibf) Combined<br>bending moment abouttwoorthogonal axes,1130Nom(10,000in\\u00b0lbf)<br>aboutoneaxis,10,170N.m(90,000in\\u00b0lbf)abouttheotheraxis(thebullgearis<br>lockedfromrotatingusingalockrackwhileloadsareapplied), torsional load5085<br>N(45T000in\\u00b0lbf)_shearload3145N(707Ibf),axialload2224N(500Ibf) Random vibration_<br>composite Glevel8.8grms Operating Temperature -40to+60\\u00b0C(-40to+140\\u00b0F)<br>107kg(235Ib) Storage >10years,operating >10years 103{'source': 'AMS_1997.pdf',<br>'page': 119}\",3,\"Chunks\"],[\"4688dd08-bd43-11ee-801f-bae7cd9d315f\",4.0638055801,10.6746759415,\"underacontrolled process, whichincluded immersion inheated cleantrichlor (43to 49\\u00b0C)ultrasonic cleaner for3to4minutes, thenair-blowing theteethandgroove. The cleaning process wasrepeated once. Thegreatest difficulty duringthedevelopment process wasthecleanliness ofthebull gearteeth.During machining operations, minute amounts ofoilbecame trapped atthe interface ofthesteelringsandthealuminum hubandseeped outduringthevacuum process, thereby contaminating thegearsurface andresulting inpoorgoldadhesion to thegearsurface. Theprocess development concentrated onthecleaning issue;for example, itwasfoundthatplacing thecleaned bullgearinthevacuum chamber and pumping downto10.4torr(totaltimeof1to1.5hr)pulledtheremaining smallamounts oftrapped oiltothesurface, whereitcouldeasilybewipedoffwithalcohol. This process wasrepeated onceortwice,thusresulting inaverycleangearsurface. Thetriodesputtering process involved placing thebullgearhorizontally onarotating tableinsideavacuum chamber withthealuminum hubcovered topandbottom (to prevent itfrombeingcoated). Thinsiliconwafercoupons, forverifying thecoating thickness, wereplaced infourplaces closetothegearteeth.Puregold(better than 99.50% purity)targets wereplaced oneachsideofthebullgear.Thevacuum chamber wasevacuated topressure ontheorderof10.5torrandback-filled withargongas.{'source': 'AMS_1997.pdf', 'page': 120}\",\"underacontrolled process, whichincluded immersion inheated cleantrichlor (43to<br>49\\u00b0C)ultrasonic cleaner for3to4minutes, thenair-blowing theteethandgroove. The<br>cleaning process wasrepeated once. Thegreatest difficulty duringthedevelopment<br>process wasthecleanliness ofthebull gearteeth.During machining operations,<br>minute amounts ofoilbecame trapped atthe interface ofthesteelringsandthealuminum<br>hubandseeped outduringthevacuum process, thereby contaminating thegearsurface<br>andresulting inpoorgoldadhesion to thegearsurface. Theprocess development<br>concentrated onthecleaning issue;for example, itwasfoundthatplacing thecleaned<br>bullgearinthevacuum chamber and pumping<br>downto10.4torr(totaltimeof1to1.5hr)pulledtheremaining smallamounts oftrapped<br>oiltothesurface, whereitcouldeasilybewipedoffwithalcohol. This process<br>wasrepeated onceortwice,thusresulting inaverycleangearsurface.<br>Thetriodesputtering process involved placing thebullgearhorizontally onarotating<br>tableinsideavacuum chamber withthealuminum hubcovered topandbottom (to prevent<br>itfrombeingcoated). Thinsiliconwafercoupons, forverifying thecoating thickness,<br>wereplaced infourplaces closetothegearteeth.Puregold(better than 99.50%<br>purity)targets wereplaced oneachsideofthebullgear.Thevacuum chamber wasevacuated<br>topressure ontheorderof10.5torrandback-filled withargongas.{'source':<br>'AMS_1997.pdf', 'page': 120}\",3,\"Chunks\"],[\"4a946534-bd43-11ee-801f-bae7cd9d315f\",3.4018511772,7.3716697693,\"0 0 0  10 61 4 49 1  20 769 615  40 962 769  60 1090, 875  80 1198 958  Additionally, it should be noted that the bearings were not run-in prior to testing.  Experimental Errors  Experimental errors arise from the errors in the measurements of thermal gradients  across the bearing and along the HFM. Experimental errors are governed by the  accuracy of the temperature sensors (L0.5 K at room temperature and 20.2 K at 20K).  Further errors result from heat loss and in the dimensions of the HFM. A summary of  experimental errors is now presented.  Room Temp: HFM calibration errors are of the order of +25%, and the dimensional  tolerances give the error in L = L2%, and the error in A = +1%.  At 48 mW heater power, typical temperature differentials across the bearing were 2 to  4K. As the sensors are only accurate to k0.5 K, experimental errors on temperature  measurements are, respectively, L50% and +25%, Combining this error with HFM  calibration and dimensional errors gives a maximum experimental error of  approximately +80 Yo.  Increasing the heater power to 180 mW resulted in larger temperature gradients, and  hence the measured temperature difference errors were reduced by a factor of 4 to 5,  Le., errors are of the order of +loo\\/& The resulting experimental errors, including HFM  calibration and dimensional errors, were of the order of +40% to f50%.  Cryoaenic Temperatures: The improved sensitivities of the temperature sensors, at  cryogenic temperatures, resulted in HFM calibration errors of +lo%. Combining these{'source': 'AMS_1996.pdf', 'page': 50}\",\"0 0 0  10 61 4 49 1  20 769 615  40 962 769  60 1090, 875  80 1198 958<br>Additionally, it should be noted that the bearings were not run-in prior to<br>testing.  Experimental Errors  Experimental errors arise from the errors in the<br>measurements of thermal gradients  across the bearing and along the HFM.<br>Experimental errors are governed by the  accuracy of the temperature sensors<br>(L0.5 K at room temperature and 20.2 K at 20K).  Further errors result from heat<br>loss and in the dimensions of the HFM. A summary of  experimental errors is now<br>presented.  Room Temp: HFM calibration errors are of the order of +25%, and the<br>dimensional  tolerances give the error in L = L2%, and the error in A = +1%.  At<br>48 mW heater power, typical temperature differentials across the bearing were 2<br>to  4K. As the sensors are only accurate to k0.5 K, experimental errors on<br>temperature  measurements are, respectively, L50% and +25%, Combining this error<br>with HFM  calibration and dimensional errors gives a maximum experimental error<br>of  approximately +80 Yo.  Increasing the heater power to 180 mW resulted in<br>larger temperature gradients, and  hence the measured temperature difference<br>errors were reduced by a factor of 4 to 5,  Le., errors are of the order of<br>+loo\\/& The resulting experimental errors, including HFM  calibration and<br>dimensional errors, were of the order of +40% to f50%.  Cryoaenic Temperatures:<br>The improved sensitivities of the temperature sensors, at  cryogenic<br>temperatures, resulted in HFM calibration errors of +lo%. Combining<br>these{'source': 'AMS_1996.pdf', 'page': 50}\",3,\"Chunks\"],[\"4ed60d32-bd43-11ee-801f-bae7cd9d315f\",4.8728985786,11.0012378693,\"QSS pe~ormed preliminary analyses on the MCF to approximate the local stresses  adjacent to the underside lip area and to determine the material properties required for  the development test articles. Two development test articles were fabricated, one from  CRES Custom 455 and one from  OSS performed early development testing using an lnstron machine to verify the  results of the preliminary analyses. The development test units were subjected to the  on-orbit loading forces of 184 N-m (250 ft-lbf) in torsion, 227 kg (550 Ibf) in axial  compression and tension, and a combined loading of 227 kg (550 Ibf) in shear and  184 N-m (250 ft-lbf) in bending. Each test unit was instrumented with four rosette  strain gauges positioned 90' apart on the inner surface of the through hole to allow for  future correlation of stresses with the detailed FEA model. In order to simulate the  loads applied by the micro conical tool tip, OSS fabricated a test jig that housed six  beryllium copper collets, using preliminary material selected for the RMCT and EMCT  collets. The MCF test units withstood all nominal loading conditions without any  evidence of yield or failure.  During an unscheduled test when the MCF test units were subjected to a maximum of  6342 N-m (8600 ft-lbf), the collets of the tool test jig yielded. Although no yielding was  observed at the MCF lip, there was substantial compressive plastic deformation from  the tool tip on the upper surface of the torque reaction shoulders.  Fracture Analvses  The MCFs do not satisfy NASA ISS requirements for a non-fracture critical component.{'source': 'AMS_1996.pdf', 'page': 384}\",\"QSS pe~ormed preliminary analyses on the MCF to approximate the local stresses<br>adjacent to the underside lip area and to determine the material properties<br>required for  the development test articles. Two development test articles were<br>fabricated, one from  CRES Custom 455 and one from  OSS performed early<br>development testing using an lnstron machine to verify the  results of the<br>preliminary analyses. The development test units were subjected to the  on-orbit<br>loading forces of 184 N-m (250 ft-lbf) in torsion, 227 kg (550 Ibf) in axial<br>compression and tension, and a combined loading of 227 kg (550 Ibf) in shear and<br>184 N-m (250 ft-lbf) in bending. Each test unit was instrumented with four<br>rosette  strain gauges positioned 90' apart on the inner surface of the through<br>hole to allow for  future correlation of stresses with the detailed FEA model.<br>In order to simulate the  loads applied by the micro conical tool tip, OSS<br>fabricated a test jig that housed six  beryllium copper collets, using<br>preliminary material selected for the RMCT and EMCT  collets. The MCF test units<br>withstood all nominal loading conditions without any  evidence of yield or<br>failure.  During an unscheduled test when the MCF test units were subjected to a<br>maximum of  6342 N-m (8600 ft-lbf), the collets of the tool test jig yielded.<br>Although no yielding was  observed at the MCF lip, there was substantial<br>compressive plastic deformation from  the tool tip on the upper surface of the<br>torque reaction shoulders.  Fracture Analvses  The MCFs do not satisfy NASA ISS<br>requirements for a non-fracture critical component.{'source': 'AMS_1996.pdf',<br>'page': 384}\",3,\"Chunks\"],[\"5168f76c-bd43-11ee-801f-bae7cd9d315f\",2.7984404564,9.5620718002,\"indicate that the elevated temperature of 100\\u00b0C was sufficient to establish a beneficial VAL between the  steel and this coating.   An effective VAL is one that inhibits high amounts of wear of the coating and the counterface over the  temperature range and relative motion experienced by a specific application. Although Ti-MoS 2 has been  previously shown to perform exceptionally well in rolling contact [4,9,12], based upon the results of these  measurements, it can be concluded that the Ti-MoS 2 coating would meet the VAL requirements better than  the tested MoS 2 and Sb 2O3\\/Au-MoS 2 coatings when in contact with reciprocating sliding steel counterfaces  over a temperature range of 30\\u00b0C to 100\\u00b0C.   It is important to point out that although the experiments were performed in laboratory air, the environment  had a very low humidity (17% RH) during the testing. Although it is expected that the wear rates of all three  coatings will increase with increasing relative humidity, undoped MoS 2 tends to experience the greatest  increase [3].  NASA\\/CP\\u20142018-219887 147{'source': 'AMS_2018.pdf', 'page': 165}\",\"indicate that the elevated temperature of 100\\u00b0C was sufficient to establish a<br>beneficial VAL between the  steel and this coating.   An effective VAL is one<br>that inhibits high amounts of wear of the coating and the counterface over the<br>temperature range and relative motion experienced by a specific application.<br>Although Ti-MoS 2 has been  previously shown to perform exceptionally well in<br>rolling contact [4,9,12], based upon the results of these  measurements, it can<br>be concluded that the Ti-MoS 2 coating would meet the VAL requirements better<br>than  the tested MoS 2 and Sb 2O3\\/Au-MoS 2 coatings when in contact with<br>reciprocating sliding steel counterfaces  over a temperature range of 30\\u00b0C to<br>100\\u00b0C.   It is important to point out that although the experiments were<br>performed in laboratory air, the environment  had a very low humidity (17% RH)<br>during the testing. Although it is expected that the wear rates of all three<br>coatings will increase with increasing relative humidity, undoped MoS 2 tends to<br>experience the greatest  increase [3].  NASA\\/CP\\u20142018-219887 147{'source':<br>'AMS_2018.pdf', 'page': 165}\",3,\"Chunks\"],[\"5aaadf8e-bd43-11ee-801f-bae7cd9d315f\",2.7705030441,9.7649517059,\"Chip qualification campaign  In parallel of the previous TRP activity, a complete qualification of the chip itself, in accordance with  space standards for the qualification of monolithic chip, is undertaken. Tests foreseen in this chip qualification activity can be divided in three steps:  \\/g131 The production control tests which occur at the manufacturer level  \\/g131 The screening tests  \\/g131 The qualification tests \\n\\nProduction control tests\\n\\nThe wafer lot acceptance tests occur at the manufacturer level: The manufacturing of the wafers is monitor and PVM data are recorded during the whole process. The final lot is inspected through scanning electron microscope. Then, in-process controls occur at the packaging level: After a pre-encapsulation visual inspection, bond strength and die shear tests are implemented. After the chip encapsulation, a dimension check of each packaged chip is performed.  Screening tests These tests are performed on all the components. Screening tests consist mainly to reject faulty chips though hard electrical and temperature testing. During those tests, the drift parameters of the chip are monitored and compare to failure criteria. The following tests are foreseen:  \\/g131 The serialized chips are subjected to a high temperature stabilisation bake (24h at 150\\u00b0C) to  determine the effect of storage at elevated temperatures without electrical stress applied. Drift parameters are monitored and control during the entire test.  \\/g131 The chips are then be subjected to several burn-in tests at 125\\u00b0C (reverse bias and power burnin) in order to eliminating marginal devices, those with inherent defects which cause time and stress dependent failures. Drift parameters are monitored and control during the entire test.  \\/g131 A measure of the drift parameters at ambient temperature is then foreseen and allow for the lot  qualification.{'source': 'AMS_2008.pdf', 'page': 197}\",\"Chip qualification campaign  In parallel of the previous TRP activity, a<br>complete qualification of the chip itself, in accordance with  space standards<br>for the qualification of monolithic chip, is undertaken. Tests foreseen in this<br>chip qualification activity can be divided in three steps:  \\/g131 The production<br>control tests which occur at the manufacturer level  \\/g131 The screening tests<br>\\/g131 The qualification tests   Production control tests  The wafer lot<br>acceptance tests occur at the manufacturer level: The manufacturing of the<br>wafers is monitor and PVM data are recorded during the whole process. The final<br>lot is inspected through scanning electron microscope. Then, in-process controls<br>occur at the packaging level: After a pre-encapsulation visual inspection, bond<br>strength and die shear tests are implemented. After the chip encapsulation, a<br>dimension check of each packaged chip is performed.  Screening tests These tests<br>are performed on all the components. Screening tests consist mainly to reject<br>faulty chips though hard electrical and temperature testing. During those tests,<br>the drift parameters of the chip are monitored and compare to failure criteria.<br>The following tests are foreseen:  \\/g131 The serialized chips are subjected to a<br>high temperature stabilisation bake (24h at 150\\u00b0C) to  determine the effect of<br>storage at elevated temperatures without electrical stress applied. Drift<br>parameters are monitored and control during the entire test.  \\/g131 The chips<br>are then be subjected to several burn-in tests at 125\\u00b0C (reverse bias and power<br>burnin) in order to eliminating marginal devices, those with inherent defects<br>which cause time and stress dependent failures. Drift parameters are monitored<br>and control during the entire test.  \\/g131 A measure of the drift parameters at<br>ambient temperature is then foreseen and allow for the lot<br>qualification.{'source': 'AMS_2008.pdf', 'page': 197}\",3,\"Chunks\"],[\"625338ee-bd43-11ee-801f-bae7cd9d315f\",3.2289829254,7.7788219452,\"curvatures  can lead to early failure,  a claim  that to the best of our understanding has never been  documented. In addition, the testing performed in this  investigation and its associated findings are of value  to designers of bearings for scanners, gimbals, and other rotary spacecraft actuators.  \\n\\n* NASA Langley Research Center, Hampton, VA   ** Fisher Aerospace, Sunnyvale, CA   + Lockheed Martin Space (retired) , Sunnyvale, CA   ++ The Aerospace Corporation, El Segundo, CA{'source': 'AMS_2020.pdf', 'page': 287}\",\"curvatures  can lead to early failure,  a claim  that to the best of our<br>understanding has never been  documented. In addition, the testing performed in<br>this  investigation and its associated findings are of value  to designers of<br>bearings for scanners, gimbals, and other rotary spacecraft actuators.    * NASA<br>Langley Research Center, Hampton, VA   ** Fisher Aerospace, Sunnyvale, CA   +<br>Lockheed Martin Space (retired) , Sunnyvale, CA   ++ The Aerospace Corporation,<br>El Segundo, CA{'source': 'AMS_2020.pdf', 'page': 287}\",3,\"Chunks\"],[\"62533a2e-bd43-11ee-801f-bae7cd9d315f\",2.8198668957,9.2502098083,\"290   Figure 4. Plot of film oil film thickness vs time during blow -off using the TFF  \\n\\nFigure 5. Plot of the data shown in Figure 4, using reciprocal of film thickness on the y -axis to show the  inverse proportionality with time  \\n\\nThese changes in viscosity and flow rate are important, because the process of resupply to the tribological  contacts may be slowed as lubricant consumption proceeds and the scarcit y of free oil leads to reduced oil  film thicknesses.  \\n\\nTribometry and Viscosity of Worn Lubricant Films  \\n\\nAnother significant reduction in oil mobility during operational use may be caused by changes in oil  composition. As the lubricant is worn in a tribological contact due to mechanical stresses and chemical  reactions, some molecules are broken into smaller components while others are polym erized into larger{'source': 'AMS_2020.pdf', 'page': 300}\",\"290   Figure 4. Plot of film oil film thickness vs time during blow -off using<br>the TFF    Figure 5. Plot of the data shown in Figure 4, using reciprocal of<br>film thickness on the y -axis to show the  inverse proportionality with time<br>These changes in viscosity and flow rate are important, because the process of<br>resupply to the tribological  contacts may be slowed as lubricant consumption<br>proceeds and the scarcit y of free oil leads to reduced oil  film thicknesses.<br>Tribometry and Viscosity of Worn Lubricant Films    Another significant<br>reduction in oil mobility during operational use may be caused by changes in oil<br>composition. As the lubricant is worn in a tribological contact due to<br>mechanical stresses and chemical  reactions, some molecules are broken into<br>smaller components while others are polym erized into larger{'source':<br>'AMS_2020.pdf', 'page': 300}\",3,\"Chunks\"],[\"661b4f52-bd43-11ee-801f-bae7cd9d315f\",3.6215083599,10.6665582657,\"resulted in incomplete removal of the flux from the solder , and this was also confirmed by radiography . This  problem  was further exacerbated by the higher revised qualification temperature. The manufacturer   proposed waveguides with electroformed flange joints, which were considered superior to the soldered  joint. After a thorough review of all options it was decided to change the waveguide configuration to the  electroformed flange joint  as proposed by the ma nufacturer . The program also decided to switch the exterior  finish to black paint inst ead of nickel plating to lower  the temperature of the middle section  during operation.  \\n\\nFigure 2. WR -34 Flexible Waveguide  \\n\\nThe flight waveguide was made up of electroformed Ni-Co flexible section with the copper flanges  attached  by the electrofor ming process. T he interior surfaces were silver plated and the exterior surfaces  were  painted with BR -127 black paint . Following a successful  batch qualification and acc eptance program, these  waveguides were installed to flight HGA assembly.  The development and flight waveguides are shown in  Figure 2.  \\n\\nWaveguide Analyses and Test s  \\n\\nThe FWG was subjected t o a comprehensive evaluation program that included testing in three phases,  development test, flight batch qualification test and acceptance tests. These test s consisted of  environmental tests and mechanical functional  tests. The condition of the FWG was monitored by  RF{'source': 'AMS_2020.pdf', 'page': 541}\",\"resulted in incomplete removal of the flux from the solder , and this was also<br>confirmed by radiography . This  problem  was further exacerbated by the higher<br>revised qualification temperature. The manufacturer   proposed waveguides with<br>electroformed flange joints, which were considered superior to the soldered<br>joint. After a thorough review of all options it was decided to change the<br>waveguide configuration to the  electroformed flange joint  as proposed by the<br>ma nufacturer . The program also decided to switch the exterior  finish to black<br>paint inst ead of nickel plating to lower  the temperature of the middle section<br>during operation.    Figure 2. WR -34 Flexible Waveguide    The flight waveguide<br>was made up of electroformed Ni-Co flexible section with the copper flanges<br>attached  by the electrofor ming process. T he interior surfaces were silver<br>plated and the exterior surfaces  were  painted with BR -127 black paint .<br>Following a successful  batch qualification and acc eptance program, these<br>waveguides were installed to flight HGA assembly.  The development and flight<br>waveguides are shown in  Figure 2.    Waveguide Analyses and Test s    The FWG<br>was subjected t o a comprehensive evaluation program that included testing in<br>three phases,  development test, flight batch qualification test and acceptance<br>tests. These test s consisted of  environmental tests and mechanical functional<br>tests. The condition of the FWG was monitored by  RF{'source': 'AMS_2020.pdf',<br>'page': 541}\",3,\"Chunks\"],[\"66d6ac5c-bd43-11ee-801f-bae7cd9d315f\",6.1382699013,10.6633863449,\"15 Lessons Learned  Several lessons were learned during the development of a deployable cover capable of holding a hermetic  seal from prototype to flight.  Hermetic Sealing  - H-seals and conflat seals are ideal for hermetically sealing interfaces, especially when metal on metal  interfaces are required.   - It is critical that the sealing surface of H-seal and knife-edge interface are pristine surfaces free of burrs,  nicks, and markings.   - Knife-edge interfaces need to be able to fully engage with the sealing surface.  - When loading H-seals onto a knife-edge, the seal needs to come down consistently perpendicular to  the surface plane. The seal cannot be rocked or unevenly engaged onto knife-edge during loading.  o For deployable mechanisms, this means compliance for the seal to come down perpendicular,  and not on an extended radius from a hinge line (if applicable).  Clampring  - When tensioning a clampring with an under-center link, the majority of loading occurs during the initial  movement of the linkage. There is limited load adjustability once the under-center linkage is near its  end of travel towards the over-center condition.  - Variability in ring ODs, even by .025 mm (.001 in), greatly affects the final tension in the clampring.  Those features need to be very tightly controlled to get repeatable results between mechanisms.   Helium Leak Testing  - It is critical to minimize all potential interfaces when testing minimal levels of Helium in the system.  - Excess background helium needs to be cleared from the immediate area of a leak detector  - The leak detector needs to be in pristine condition to measure at noise floor. Any undesirable{'source': 'AMS_2022.pdf', 'page': 29}\",\"15 Lessons Learned  Several lessons were learned during the development of a<br>deployable cover capable of holding a hermetic  seal from prototype to flight.<br>Hermetic Sealing  - H-seals and conflat seals are ideal for hermetically sealing<br>interfaces, especially when metal on metal  interfaces are required.   - It is<br>critical that the sealing surface of H-seal and knife-edge interface are<br>pristine surfaces free of burrs,  nicks, and markings.   - Knife-edge interfaces<br>need to be able to fully engage with the sealing surface.  - When loading<br>H-seals onto a knife-edge, the seal needs to come down consistently<br>perpendicular to  the surface plane. The seal cannot be rocked or unevenly<br>engaged onto knife-edge during loading.  o For deployable mechanisms, this means<br>compliance for the seal to come down perpendicular,  and not on an extended<br>radius from a hinge line (if applicable).  Clampring  - When tensioning a<br>clampring with an under-center link, the majority of loading occurs during the<br>initial  movement of the linkage. There is limited load adjustability once the<br>under-center linkage is near its  end of travel towards the over-center<br>condition.  - Variability in ring ODs, even by .025 mm (.001 in), greatly<br>affects the final tension in the clampring.  Those features need to be very<br>tightly controlled to get repeatable results between mechanisms.   Helium Leak<br>Testing  - It is critical to minimize all potential interfaces when testing<br>minimal levels of Helium in the system.  - Excess background helium needs to be<br>cleared from the immediate area of a leak detector  - The leak detector needs to<br>be in pristine condition to measure at noise floor. Any undesirable{'source':<br>'AMS_2022.pdf', 'page': 29}\",3,\"Chunks\"],[\"6772cbf0-bd43-11ee-801f-bae7cd9d315f\",2.8167743683,9.3256397247,\"heating and from austentie to a mix of martensite and R-phase upon cooling, the test could not be  considered repeatable, and the predicate of the model is violated. While the model strongly correlates with  physical principles and expected actuation behavior, training SMA to transition from only detwinned  martensite to austenite at a specific temperature is non-trivial.   Methods exist to remove the spurious phases, though more testing is warranted on purchased specimens  with consistent material properties. An increase in the initial furnace temperature leads to a shorter and  less pronounced R-phase plateau [12]. In Halvorson et al [10], spike phases were postulated as effects of  quenching in liquid nitrogen instead of an ice bath; this has been determined to be incorrect. Multiple DSC  test iterations with an ice water quench resulted in spike phases occurring where R-phase is expected near  the Austenite finish temperature. The evolution to R-phase to glass phase is poorly documented in  literature. It is asserted that the spike phases are glass transition phases; glass phases can occur ranging  from 12%-82% nickel by mass [13].   Predictive Actuation Model  The thermal, constitutive, and kinematic behavior of the SMA actuation process was modeled with a  MATLAB simulation code divided into three elements: a heat transfer model, a thermo-mechanical model,  and a kinematic model. SMA transient thermal response to PH operation in the bending region was  determined using a quasi-3D, finite-difference heat transfer model with simulated conduction, convection,  and radiation heat transfer effects corresponding to lab and PH input conditions. The thermo-mechanical{'source': 'AMS_2022.pdf', 'page': 78}\",\"heating and from austentie to a mix of martensite and R-phase upon cooling, the<br>test could not be  considered repeatable, and the predicate of the model is<br>violated. While the model strongly correlates with  physical principles and<br>expected actuation behavior, training SMA to transition from only detwinned<br>martensite to austenite at a specific temperature is non-trivial.   Methods<br>exist to remove the spurious phases, though more testing is warranted on<br>purchased specimens  with consistent material properties. An increase in the<br>initial furnace temperature leads to a shorter and  less pronounced R-phase<br>plateau [12]. In Halvorson et al [10], spike phases were postulated as effects<br>of  quenching in liquid nitrogen instead of an ice bath; this has been<br>determined to be incorrect. Multiple DSC  test iterations with an ice water<br>quench resulted in spike phases occurring where R-phase is expected near  the<br>Austenite finish temperature. The evolution to R-phase to glass phase is poorly<br>documented in  literature. It is asserted that the spike phases are glass<br>transition phases; glass phases can occur ranging  from 12%-82% nickel by mass<br>[13].   Predictive Actuation Model  The thermal, constitutive, and kinematic<br>behavior of the SMA actuation process was modeled with a  MATLAB simulation code<br>divided into three elements: a heat transfer model, a thermo-mechanical model,<br>and a kinematic model. SMA transient thermal response to PH operation in the<br>bending region was  determined using a quasi-3D, finite-difference heat transfer<br>model with simulated conduction, convection,  and radiation heat transfer<br>effects corresponding to lab and PH input conditions. The thermo-<br>mechanical{'source': 'AMS_2022.pdf', 'page': 78}\",3,\"Chunks\"],[\"f723a572-bd42-11ee-801f-bae7cd9d315f\",5.6985211372,10.8252334595,\"Parallel Corrugated  Diaphragms  m  I  s 7 1  I k- Bearings  Flex link  L Interface  r \\/- Titanium Housing  Figure 4. CRISM Diaphragm Bearing Assembly Anti-Sunward  Radiator  The CRlSM duplex bearing pairs were separated by titanium spacers so that the preload offset would  remain constant over temperature. However, the difference between the 440C inner and outer rings and  the titanium shaft and housing resulted in a reduction of clearance as temperature decreased (Figure 6).  A reduction of clearance results in a decrease of the contact angle. However, the bearings are only going  to experience substantial axial loads during the launch. The bearings were tested to -196\\u00b0C and  continued to rotate freely. The few disadvantages of preloading are more than offset by the following  advantages:  Reduces axial and radial runout of the rotating shaft. Required for the encoder disk to read head  alignment  Reduces the shaft deflection under load and improves its assembled stiffness  Removes free play in the bearing set, keeping the bearing set loaded in-order to avoid skidding of  the balls  Minimizes the peak stresses that occur during the maximum loading events by ensuring the load  on the bearings is shared by more balls in each bearing 0  Decreases bearing noise  In addition to the axial preload, the CRlSM bearings employed a light interference fit, 12.7 pm (0.0005 in)  in the bearingkhaft fit and 15.2 pm (0.0006 in) in the bearing\\/housing fit.  14{'source': 'AMS_2006.pdf', 'page': 28}\",\"Parallel Corrugated  Diaphragms  m  I  s 7 1  I k- Bearings  Flex link  L<br>Interface  r \\/- Titanium Housing  Figure 4. CRISM Diaphragm Bearing Assembly<br>Anti-Sunward  Radiator  The CRlSM duplex bearing pairs were separated by<br>titanium spacers so that the preload offset would  remain constant over<br>temperature. However, the difference between the 440C inner and outer rings and<br>the titanium shaft and housing resulted in a reduction of clearance as<br>temperature decreased (Figure 6).  A reduction of clearance results in a<br>decrease of the contact angle. However, the bearings are only going  to<br>experience substantial axial loads during the launch. The bearings were tested<br>to -196\\u00b0C and  continued to rotate freely. The few disadvantages of preloading<br>are more than offset by the following  advantages:  Reduces axial and radial<br>runout of the rotating shaft. Required for the encoder disk to read head<br>alignment  Reduces the shaft deflection under load and improves its assembled<br>stiffness  Removes free play in the bearing set, keeping the bearing set loaded<br>in-order to avoid skidding of  the balls  Minimizes the peak stresses that occur<br>during the maximum loading events by ensuring the load  on the bearings is<br>shared by more balls in each bearing 0  Decreases bearing noise  In addition to<br>the axial preload, the CRlSM bearings employed a light interference fit, 12.7 pm<br>(0.0005 in)  in the bearingkhaft fit and 15.2 pm (0.0006 in) in the<br>bearing\\/housing fit.  14{'source': 'AMS_2006.pdf', 'page': 28}\",3,\"Chunks\"],[\"f976480c-bd42-11ee-801f-bae7cd9d315f\",7.2174172401,7.0499854088,\"Nose Landina Gear Udock Mechanism  The Space Shuttle Orbiter\\u2019s nose landing gear, nose landing gear door, and nose landing gear uplock  mechanism, shown in Figure 1, are interconnected, and must be rigged and operated together.  Following replacement of the door environmental seal, rigging was performed to achieve proper seal  compression. During nose landing gear cycling, the gear uplock indication did not illuminate because the  mechanism did not reach the full uplock condition. Binding in the rotational fitting between the uplock  fitting and the bellcrank prevented the uplock mechanism from going to the full over-center position for  gear uplock. Measurements of the width of the bellcrank and the internal width of the fitting showed an  interference fit between the two assemblies. Rework on the bushings per specification requirements  removed the interference condition, allowing the bellcrank to move freely.  Lesson Learned: Proper tolerancing and inspection are critical to preventing interferences in  mechanical systems.  Bungee location,  not shown  Environmental door seal,  both doors Shock strut  Fr\\/  Figure 1. Nose Landing Gear Mechanisms  During subsequent nose landing gear retract operations, there was an early indication that the gear uplock  mechanism was in the gear-up position. As the shock strut was entering the wheel well and bringing the  doors closed, the gear stalled prior to being fully up and locked. After an immediate halt to operations the  gear fell freely to the down position. It was observed that the uplock hook was in the gear-up position, thus  preventing the uplock roller from engaging. Upon investigation, it was discovered that when hydraulic{'source': 'AMS_2006.pdf', 'page': 128}\",\"Nose Landina Gear Udock Mechanism  The Space Shuttle Orbiter\\u2019s nose landing<br>gear, nose landing gear door, and nose landing gear uplock  mechanism, shown in<br>Figure 1, are interconnected, and must be rigged and operated together.<br>Following replacement of the door environmental seal, rigging was performed to<br>achieve proper seal  compression. During nose landing gear cycling, the gear<br>uplock indication did not illuminate because the  mechanism did not reach the<br>full uplock condition. Binding in the rotational fitting between the uplock<br>fitting and the bellcrank prevented the uplock mechanism from going to the full<br>over-center position for  gear uplock. Measurements of the width of the<br>bellcrank and the internal width of the fitting showed an  interference fit<br>between the two assemblies. Rework on the bushings per specification<br>requirements  removed the interference condition, allowing the bellcrank to move<br>freely.  Lesson Learned: Proper tolerancing and inspection are critical to<br>preventing interferences in  mechanical systems.  Bungee location,  not shown<br>Environmental door seal,  both doors Shock strut  Fr\\/  Figure 1. Nose Landing<br>Gear Mechanisms  During subsequent nose landing gear retract operations, there<br>was an early indication that the gear uplock  mechanism was in the gear-up<br>position. As the shock strut was entering the wheel well and bringing the  doors<br>closed, the gear stalled prior to being fully up and locked. After an immediate<br>halt to operations the  gear fell freely to the down position. It was observed<br>that the uplock hook was in the gear-up position, thus  preventing the uplock<br>roller from engaging. Upon investigation, it was discovered that when<br>hydraulic{'source': 'AMS_2006.pdf', 'page': 128}\",3,\"Chunks\"],[\"f9764b4a-bd42-11ee-801f-bae7cd9d315f\",3.5726897717,6.7969379425,\"The SRS's of the shock pulses are shown in Figure 6 and the time histories of the pulses are shown in  Figure 7.  The response of the payload mass was measured with accelerometers at the locations shown in Figure5.  The peak response of the mass was measured at the channel 8 (in axis) accelerometer location. The  peak response from the shock pulse was peak input value of 805 g's and a peak response of the payload  mass was 21.2 g's. This was a reduction of 31 dB. The amount of isolation due to base shock input can  be seen by viewing the base input time history on the same plot as the mass response as plotted in  Figure 7.  The SRS of the 18.1-kg (40-lb) mass to both of the shock pulses is shown in Figure 6. The SRS of the  response reflects the fact that, in the higher frequencies, the first input shock pulse is higher and this is  reflected in the response of the mass. One noticeable artifact of the mass response in Figure 6 is that  there is a peak between 560 Hz and 630 Hz even though the break frequency of both pulses is around  800 Hz. The 630 Hz peak is very close to known surge frequency of the isolator main springs. The sine  vibration data only goes to 2 kHz; therefore, the correspondence to the peaks in the SRS data can only  be tracked to 2 kHz.  Conclusions about the Shock Beam Test Results  The following conclusions can be made about the test results.  0 The shock beam test was able to achieve the input levels required by the potential isolation  system.  The isolators in the bipod configuration were able to decrease mass responses relative to the  maximum input by 30 dB or more. These Isolators eat shock!{'source': 'AMS_2006.pdf', 'page': 159}\",\"The SRS's of the shock pulses are shown in Figure 6 and the time histories of<br>the pulses are shown in  Figure 7.  The response of the payload mass was<br>measured with accelerometers at the locations shown in Figure5.  The peak<br>response of the mass was measured at the channel 8 (in axis) accelerometer<br>location. The  peak response from the shock pulse was peak input value of 805<br>g's and a peak response of the payload  mass was 21.2 g's. This was a reduction<br>of 31 dB. The amount of isolation due to base shock input can  be seen by<br>viewing the base input time history on the same plot as the mass response as<br>plotted in  Figure 7.  The SRS of the 18.1-kg (40-lb) mass to both of the shock<br>pulses is shown in Figure 6. The SRS of the  response reflects the fact that, in<br>the higher frequencies, the first input shock pulse is higher and this is<br>reflected in the response of the mass. One noticeable artifact of the mass<br>response in Figure 6 is that  there is a peak between 560 Hz and 630 Hz even<br>though the break frequency of both pulses is around  800 Hz. The 630 Hz peak is<br>very close to known surge frequency of the isolator main springs. The sine<br>vibration data only goes to 2 kHz; therefore, the correspondence to the peaks in<br>the SRS data can only  be tracked to 2 kHz.  Conclusions about the Shock Beam<br>Test Results  The following conclusions can be made about the test results.  0<br>The shock beam test was able to achieve the input levels required by the<br>potential isolation  system.  The isolators in the bipod configuration were able<br>to decrease mass responses relative to the  maximum input by 30 dB or more.<br>These Isolators eat shock!{'source': 'AMS_2006.pdf', 'page': 159}\",3,\"Chunks\"],[\"fa0b2d96-bd42-11ee-801f-bae7cd9d315f\",7.4568123817,7.3331599236,\"tool and onto the aft bulkhead of the spacecraft and not the SSRD. Figure 4. SSRD Torque Retention Tool  A key feature of the SSRD is it is a mechanically redundant device. It achieves this redundancy because  only one of the split spool halves needs to move laterally in order for the rod end to be released.  However, this is only true as long as the rod end is not constrained from exiting the unit. In order to  maintain mechanical redundancy at the system level, the retraction system needed to be designed such  152{'source': 'AMS_2006.pdf', 'page': 166}\",\"tool and onto the aft bulkhead of the spacecraft and not the SSRD. Figure 4.<br>SSRD Torque Retention Tool  A key feature of the SSRD is it is a mechanically<br>redundant device. It achieves this redundancy because  only one of the split<br>spool halves needs to move laterally in order for the rod end to be released.<br>However, this is only true as long as the rod end is not constrained from<br>exiting the unit. In order to  maintain mechanical redundancy at the system<br>level, the retraction system needed to be designed such  152{'source':<br>'AMS_2006.pdf', 'page': 166}\",3,\"Chunks\"],[\"fc8c2c96-bd42-11ee-801f-bae7cd9d315f\",4.3813810349,5.9784049988,\"*  Large Scale Measurement Device >  *\\/ Pane's  Base  # ,\\/- Frame  d . Base  (41 Frame  Laser Tracker  + Laser Tracker Targets  Figure 8. SAR Panel Location Measurement Setup  The line of sight to these points in the test range is limited due to mechanisms and guardrails in the back  of the panels. Different set-ups were required. Figure 8 shows a typical set-up schematic.  The measured point positions were finally given in the mechanical build (mb) coordinate system that was  predefined. Repeatability of the measurement was verified and the error from nominal position was  calculated. The final Laser tracker measurement results of the three-dimensional positional errors are  listed in Table 2. The error is defined as the difference between nominal and measured positions. The  maximum error calculated is 0.61 mm (0.024 inch) in the X direction. The compound error is attributed to  the tolerance of machining and assembly, the relative position of each panel to the reference frame and  the measurement set-up accuracy.  340{'source': 'AMS_2006.pdf', 'page': 354}\",\"*  Large Scale Measurement Device >  *\\/ Pane's  Base  # ,\\/- Frame  d . Base  (41<br>Frame  Laser Tracker  + Laser Tracker Targets  Figure 8. SAR Panel Location<br>Measurement Setup  The line of sight to these points in the test range is<br>limited due to mechanisms and guardrails in the back  of the panels. Different<br>set-ups were required. Figure 8 shows a typical set-up schematic.  The measured<br>point positions were finally given in the mechanical build (mb) coordinate<br>system that was  predefined. Repeatability of the measurement was verified and<br>the error from nominal position was  calculated. The final Laser tracker<br>measurement results of the three-dimensional positional errors are  listed in<br>Table 2. The error is defined as the difference between nominal and measured<br>positions. The  maximum error calculated is 0.61 mm (0.024 inch) in the X<br>direction. The compound error is attributed to  the tolerance of machining and<br>assembly, the relative position of each panel to the reference frame and  the<br>measurement set-up accuracy.  340{'source': 'AMS_2006.pdf', 'page': 354}\",3,\"Chunks\"],[\"ffe5f9f8-bd42-11ee-801f-bae7cd9d315f\",6.9942522049,9.5941085815,\"same length and are stowed coincident with each other. The latches fit in the annular gap between adjacent tubes in a stiffening ring at the lower end of each tube. The adjacent larger tube in turn necks down to a thin stiffening ring at the upper end. The stiffening ring helps to center and align the adjacent smaller tube and to lessen local deformations between the latched segments in bending.  Tube Latching  Small tapered pins are distributed circumferentially in the stiffening ring at the lower end of each tube. The pins are loaded radially outward by short springs to engage with tapered holes at the upper end of each larger adjacent tube, as shown in Figure 4. When stowed, the springs and pins are compressed by the interior surface of the adjacent larger tube. During  Figure 4.  Tapered pins used for latching{'source': 'AMS_2012.pdf', 'page': 143}\",\"same length and are stowed coincident with each other. The latches fit in the<br>annular gap between adjacent tubes in a stiffening ring at the lower end of each<br>tube. The adjacent larger tube in turn necks down to a thin stiffening ring at<br>the upper end. The stiffening ring helps to center and align the adjacent<br>smaller tube and to lessen local deformations between the latched segments in<br>bending.  Tube Latching  Small tapered pins are distributed circumferentially in<br>the stiffening ring at the lower end of each tube. The pins are loaded radially<br>outward by short springs to engage with tapered holes at the upper end of each<br>larger adjacent tube, as shown in Figure 4. When stowed, the springs and pins<br>are compressed by the interior surface of the adjacent larger tube. During<br>Figure 4.  Tapered pins used for latching{'source': 'AMS_2012.pdf', 'page': 143}\",3,\"Chunks\"]]}"
+}
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