ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
288   Figure 12.  Geared Motor Output Torque vs Speed for Three Motors  

Environmental Induced Effects on Gear Output Torque    The environmental effect s on the geared motor output torque are numerous. In the scope of this paper,  only the temperature effect is evaluated. Applying Equation 14, the geared motor output torque versus  operating speed at three operating temperatures: 2 43 oK (cold), 293oK (RT), and 348oK (hot) and the given  applied current of 1 a mpere is plotted in Figure 13. From Figure 13, at the nominal operating speed of   461 rpm, the output torque variation can be as high as 75%.    It is observed that the combined effect of the motor paramet er variation and the temperature can cause the  output torque variation to be as much as 82%. This is a conservative prediction due to the small motor  sample size (i.e., 3) and the motor parameters are evaluated for the primary winding and the deployment direction only.
289   Figure 13.  Geared Motor Output Torque vs Speed for CDA -8124 at Different   Operating Temperature  

Geared Motor and Gearbox Torque Efficiency Calculation      For the motor evaluation, it’s useful to show the geared motor and the gearbox torque efficiency. In general,  the torque eﬃciency η of a motor is deﬁned as the ratio between the input electrical power, i.e., the product  of voltage times current, and the output mechanical power minus the power loss due to mechanical means, which is expressed as follows:  

 η𝑣𝑣𝑚𝑚𝑅𝑅𝑚𝑚𝑚𝑚=(𝑘𝑘𝑇𝑇𝑖𝑖−𝑇𝑇𝑙𝑙𝑚𝑚𝑠𝑠𝑠𝑠)(1−𝑘𝑘𝑠𝑠𝑣𝑣) 𝑘𝑘𝑇𝑇𝑖𝑖   (15)  where T loss is the net projected torque loss of the motor and the gearbox to the motor rotor location due to  speed, temperature, and windage, and n gslide is the gearbox tooth power efficiency due to applied torque.    Using Equation 2 with motor parameters from Figure 11, t he geared motor torque efficiency can be  calculated as follows:          (16)
290 Figure 14 shows the geared motor torque efficiency as a function of applied current and geared motor  output torque at room temperature. From Equation 16, it is clear that the mechanical efficiency for the  geared motor is a strong function of several motor operating parameters such as motor transmitted torque,  operating speed, operating temperature, humidity, atmospheric pressure, etc. Therefore, it is not very  specific and should be used as a guideline to motor selection or design/configuration  comparison.  

Figure 1 4.  Geared Motor Torque Efficiency vs Motor Rotor Speed at Different Temperature s 

 The gearbox torque  efficiency can be calculated as follows:             (17)    Where the values for each parameters for each motor are defined and can be found in Figure 11.
291   Figure 15.  Simulated Power Efficiency vs. Motor Rotor Speed and Operating Temperature   at 1-A Applied Current 

It is observed that at low input torque, the gearbox torque efficiency is dominated by the Coulomb, viscous, and detent torques. However, at close to full load, the gearbox efficiency is controlled by gear tooth sliding friction coefficient.  

Conclusion    A brushless DC motor torque model has been developed in order to extract  the geared motor torque  component loss for the gearbox and the motor using data from motor measurement s and the geared motor  testing. The geared motor output torque and mechanical effi ciency vary from motor to motor and are a  strong function of the transmitted torque, rotor speed, and operating temperature.  

If the geared output torque requirement is stringent,  the motor characterisation process and the selected  operating condition as shown above are a must in order to reduce the uncertainty.  

Lessons Learned  

1. The geared motor output torque significantly varies from motor to motor and with operating conditions .  The output torque variation has been shown to be 82%. If additional factors are accounted for, it can  be 100% or higher . By characterizing each motor and selecting optimal operating conditions, the output  variation can be significantly reduced.   2. The geared motor tests at the motor vendor should be well defined and the results shoul d be verified  at each step to assure that the physics can be understood or explained.   3. The MPB viscous torque model shows a good correlation with the geared motor test data and is  proportional to the speed to the power of 0.667. The motor bearing Coulomb torque can be reasonably  predicted using the ball bearing torque model in BRGS10C with a friction coefficient of 0.15 to 0.20 for small bearings (OD < 1 in (2.5 cm) )  4. The gearbox efficiency is not a constant and a non-linear function throughout operating conditions .
292 References  

1. Ward, Peter, “ Effects of Bearing Cleaning and Lube Environment on Bearing Performance”, 29th  Aerospace Mechanisms Symposium Proceeding, Houston, TX, 1995.   2. Leveille, A., The Aerospace Corporation Bearing Analysis Program, BRGS10C version 10C, 2005   3. Tran, A.N., Halpin, J.D.,”  JAMES WEBB SPACE TELESCOPE DEPLOYMENT TOWER ASSEMBLY  DEPLOYING ANOMALY AND  LESSONS LEARNED”, 16th European Space Mechanisms and  Tribology Symposium 2015, Bilbao, Spain -  23rd to 25th September 2015   4. Vrancik, J.E., “Prediction of Windage Power Loss In Alternators”, NASA TN D -4849, Lewis Research  Center, Cleveland, Ohio, 1968   5. Perso nal communication with staff at  CDA  InterCorp,  Florida, USA  Dec 23, 2015   6. Brushless Permanent Magnet Motor Engineering Reference Data, CDA InterCorp Catalog  

Acknowledgement 

The author would like to thank Paul Reynolds, Optical Telescope Elements Mechanisms IPT leader, Dr.  Scott Texter, NGC JWST Optical Telescope Elements program manager for their guidance, and my NGC  colleagues Shaun O’Neill and Stan Klyza for technical discussion. The technical discussion with CDA  InterCorp is also  appreciated.
293 Development of a  Deployment and Latching Mechanism for a Pointing System    on the Neutron Star Interior Composition Explorer (NICER)   for Goddard Space Flight Center  

Robert H. Berning* and John J. Holzinger * 

Abstract    Many space applications require high stiffness and stable  latching systems . The Deployment and Latching  Mechanism design for the Neutron Star Interior Composition Explorer (NICER), provides a stable, highstiffness structure  (> 1,500, 000 in- lb/rad  (169 MN -m/rad) ) in the  deployed and stow ed positions . The  Deployment  and Latching Mechanism consists of a rotar y deployment actuator,  rotary latching actuator and  an over-center slider crank mechanism latch. This design allows the NICER system  to be repetitively  latch ed in both stowed and deployed positions , providing locking (high stiffness) in both rotational directions .  This paper describes the development, design, incorporation of lessons learned and testing of the  Deployment and Latching Mechanism . 

Figure 1.  Neutron Star Interior Composition Explorer (NICER)   (shown i n Deployed and Latched Position)  

                                                     * Moog Inc.  Chatsworth, CA  

Proceedings of the 43rd Aerospace Mechanisms Symposium, NASA Ames Research Center, May 4-6, 2016   NICER Instrument   Space Station  (FRAM)  Deployment and  Positioning System   (DAPS)  Deployment and  Latching Mechanism   (Subassembly of DAPS)
294 Introduction  

The deployment and latching m echanism is part of the Deployment and Positioning System (DAPS) . The  DAPS is an articulating mechanis m for the NICER program consisting  of a deployment and latching   mechanism, a pointing mechanism , and an interconnecting boom structure. The D APS provides a stable,  high-stiffness structure in the deployed and latched positions for a star -tracker system that “enables  (together with NICER’s GPS -based absolute timing) high- precision pulsar light -curve measurem ents  through ultra- deep exposures spanning the 18- month mission lifetime.”1 It has a deployment articulating  arm that needs to be deployed 84 ° from the vertically stowed position and latched prior to pointing  mechanism activation. The arm and the full instrument is  required  to be stowed and latched multiple times  on board the International Space Station (ISS). A final stow and latch cycle is required for safe removal  from the Space Station. DAPS has specially designed interfaces at the ISS interface and at the payload .  The latch can provide locking for the ride to the ISS, multiple latched stow and deploy cycles on the ISS,  and if required, it will perform a final stow and latch for removal from the ISS if necessary.  

                   Figure 2.  Deployment and Positioning System (DAPS)   (shown i n Stowed  and Latched Position)  

Deployment and Latching Mechanism Development    The deployment and latching mechanism is a subassembly of the NICER DAPS  located between the Space  Station Flight Releasable Attachment Mechanism ( FRAM ) interface and the boom arm with a dual axis  pointing gimbal  attached at the boom end that is the x-ray instrument attach point . The deployment and  latch mechanism enables positioning and latching of the NICER instrument in both stowed and deployed  positio ns. It consists of a bracket that is mounted onto the FRAM and employs a Moog deployment actuator ,  a Moog latching actuator , and the latching mechanism . The deployment mechanism is operated by these  two electro- mechanical actuators, each having a stepper m otor coupled to a zero- backlash gear  transmission. The boom is attached to the deployment actuator . This type of actuator configuration is a very  common approach for many space- related deployment applications, thus, there is extensive heritage  justifying t he use of these actuation mechanisms .    The deployment actuator travel from stop to stop is 84°. At both the deployment actuator’s sto wed and  deployed positions , the latching actuator  engages the latch to hold the deployment actuator position. Prior  to deployment actuator operation , the latch actuator releases the lat ch to allow the d eployment actuator to  stow or deploy . This is accomplished using an over -center  slider crank mechanism that prevents  the  Deployment and  Latching Mechanism  Dual Axis Pointing  Gimbal   Deployment and  Latching Mechanism  Boom Assembly  NICER Instrument  Interface   Space Station  (FRAM)  Interface
295 deployment actuator from rotating  in the stowed and deployed  position s. In one direction,  the d eployment  actuator hardstop prevents movement  and in the other direction,  movement  is prevented by the over-center   slider crank latch mechanism contacting the latching actuator hardstop.  

The initial idea was to develop a latch that utilized physical hardstops for stowed and deployed positions  without using motor power to hold position. The NICER program required two position high stiffness latch  locations capable of reacting to loads in both CW and CCW directions . Heritage designs utilized either  motor holding capabilities, additional latch actuators for each position, or an over-center  device that held  one position but was not a high- stiffness mechanism . Based on program requirements, a new latc h  mechanism would be needed to do the job.  

Moog  quickly ascertained that the demanding stiffness requirement s of this application would not be  achieved with a simple dual -sided wedge design, and instead focused our design concept around an  efficient single- sided wedge that could take advantage of the integrated hardstop.  The latching concept  was completed with an over-center  crank, resulting in a two- direction latch with physical stops in both  directions. The latching actuator provides the crank portion of the slider crank mechanism, positioning the  plunger wedge in the latched or released position, while a bolt w ith integral wedge works as the slider . An  integral travel stop was added to the bolt assembly to limit deployment actuator travel when non -operation  loads are applied. Mechanism design locks latch in place by establishing a ridged structure between the  deployment actuator and the latching actuator hardstop. Under high torque loading,  the latch design will  not open , which prevents gapping or unlatching of the mechanism.  

Figure 3.  Latch System Concept   Though Moog heritage designs included Extra- Vehicular Activity (EVA) bolts, a unique addition to the  NICER DAPS actuator mechanisms is the integration of the Extra Vehicular Activity (EVA) with the Extra– Vehicular Robot (EVR) attachment that is designed to provide the manual override of the actuator in case  of loss of power . This EVA/EVR interface is designed to be operated by the astronauts or by the ISS robotic  manipulator arm in a zero- gravity environment using a standard electrical powered driver . A magnetic clutch  is provi ded to protect the zero- backlash gear transmission from being over -stressed . Also, the operating  nomenclature on the actuator provides clear instruction and ease of viewing to reduce the potential for misuse . Each actuator’s output position is indicated by  a non- redundant potentiometer . NASA was planning  to use cameras located on the robotic arm and possibly an astronaut’s eyes to indicate revolutions of the
296 EVA/EVR bolt from hardstops or home positions to “ease” into the hardstops by counting revolutions a nd  starting and stopping when hardstops were about to be engaged.  

The latch design must prevent move ment of the deployment actuator  during NICER operation . Knowing  this, a detailed free body diagram was created and used to create a mathematical model of the latch mechanism . The model  optimiz es the latch mechanism, by balancing input load with latch component loads  and latching actuator performance. It also calculates all loads within the latch mechanism , torsional loading  of the deployment actuator, latch  component loads, friction, etc . The model can resize the latch mechanism  for alternate load environments . 

Figure 4.  Latch System Free Body Diagram    Stiffness of the overall deployment system when latched was a critical design criteria . Each component  was designed to maximize its  individual stiffness  to assure the latching mechanisms combined stiffness is   as high as possible. A detailed tolerance stack was used to assure no misalignments or binding w ere  possible.    Once the design was complete, a high- fidelity proof of  concept tool was fabricated. The proof of concept   tool was used to develop the assembly process and verify performance. Math model calculations were  confirmed with t he proof of concept tool , which correctly calculates the latch force and latch actuator torque.    To address life for the deployment and latching mechanism, the possibility of galling, seizing, friction and  wear was reduced or eliminated. Life of the deployment and latching mechanism is accomplished with a  four-fold approach.   1. All sliding surfaces, including shims, bolt  components , slider block, link and link  pins are made from  non-galling stainless steel or a super alloy that is not galling coupled w ith the non -galling steel.   2. All sliding s urfaces  in the latching mechanism are diffusion layer hardened. Internally these parts  retain the toughness of the original material and hardens  only a thin outer layer . This hardening  further protects the sliding surfaces from galling.   3. To help minimize fr iction of the sliding surfaces, protect these surfaces from galling and seizing,  and reduc e wear, a dry film lubricant is  applied. This low temperature impingement -applied  lubricant also is enhanced with the application of other lubricants (co- lubrication) .   4. To further minimize the effects of friction, chance of galling and seizing, and wearing, a secondary grease lubrication was applied to all sliding surfaces of the deployment and latching mechanism.
297   Figure 5.  Proof of concept tooling  

The proof of concept tool utilized  flight latching mechanism parts but used representative output flanges for  both actuators and a plate instead of the flight  mounting bracket . Duplex pair bearings were also mounted  under each of  the output flanges to simulate the flight configuration. The goal was to check the mechanism’s  tolerances, required torques and the torsional stiffness of the mechanism’s deployment actuator at the stowed and deployed positions . The pro of of concept tool also enabled assembly, shimming,  and load  setting operations to be finalized prior to the flight unit build.    Proof of concept testing was very successful, the latching mechanism exceeded all program requirements .  Flight latch assembly began after complet ion of latch tool proof of concept testing. In-process stiffness  testing reconfirmed math model results and verified flight latch stiffness.  

              Figure 6.  Flight Latching Mec hanism   Deployment Actuator   Latching  Actuator  Latch Mechanism  EVA/EVR Bolt   (Covered by Dust Cap)
298 The Flight unit successfully completed in- process stiffness testing, with only a slight re duction of overall  stiffness . Deployment and latch assembly was then completed with no issues . The deployment and latch  mechanism successfully co mpleted protoqual  testing with no issues . Latch performance did not change  after exposure to vibration and thermal vacuum.  

Proof -Of-Concept Model Stiffness  

NICER Deployment and Latching Mechanism  Stiffness  

Figure 7.  Stiffness results  

215,800 N -m/rad   195,237 N-m/rad
299   Life Test Unit Initial Stiffness  

Life Test Unit Final Stiffness  

Figure 8.  Life stiffness results  

Life testing was completed using the proof of concept tooling . All non -flight latch components were replaced  with flight versions and the latch was cycled 3000 cycles (latch then unlatch), 3 000 in both deployed and  stowed position.  No wear or damage was observed, and latch performance was unchanged . Life testing  demonstrated the effectiveness of the design and processes to greater than 2X the NICER life requirement .  Extended life testing is currently in process.   325,283 N-m/rad   324,040  N-m/rad
303 Tape Hinge/Lenticular Strut Hinge Qualification and Evolution  

Donald Gibbons * 

Abstract    A lenticular strut/carpenters tape hinge was developed, qualified, and flown, with successful operation on  orbit on several vehicles . The lessons from this effort were incorporated into a second- generation design .  This design is suitable for large lightweight structures such as thermal shields and solar sails . While the  initial design relied on the actual gossamer structure for the hinge functional testing, the evolved design  can be tested in component -size facilities without loss of test fidelity, and integrated into various assemblies  at vehicle integration resulting in a more affordable hinge,  suitable for lightweight mechanisms.  

Introduction  

The hinges have been built and were qualified and acceptance tested to a typical set of  spacecraft  mechanisms  test requirements:  torque output, vibration testing, thermal vacuum cycling and operation,  and for the prototype or qualification unit , life testing. This hinge is not intended for high torque applications,  but is suitable for lightweight deployable structures such as thermal shields or baffles, or for solar sail applications.  

The hinge is shown in Figure 1. This simple design contains few moving parts . Functioning as a hinge, the  basic mechanism is a flexure. This makes it ideal for lightweight applications bec ause of the inherent  insensitivity to temperature extremes during operation . The iterations on this design from qualification to  flight hardware to re- use have improved the processes and decreased the costs associated with the build  and test of the hinge.  

Figure 1.  Stowed Installed Hinge                                                        * Lockheed Martin Space Systems Company, Sunnyvale, CA  

Proceedings of the 43rd Aerospace Mechanisms Symposium, NASA Ames Research Center, May 4-6, 2016
305 The stowed radius, R 2, has a natural curvature as shown in Figure 4. For this installation , the hinge support  and frame control the stowed radius  to a smaller radius than the natural curvature due to the interface  constraints . The smaller stowed radius naturally results in higher stresses, but the compliance  of the frame  allowed some movement of the ends of the hinge, resulting in a larger radius and lower stresses in the  hinges than a rigid  structure would allow . An interesting result of this stowed configuration is that the hinge  will move laterally as well in rotation to due to this tighter curvature and additional strain energy . This effect  is shown in Figure 6. While this doesn't impact the operation of the hinge, the additional movement of the  mechanism should be considered in the deployment dynamics and clearances during the release of the  mechanism.  

Figure 2.  Folding Hinge  

Figure 3.  Detail of Secondary Buckle and Hinge Support  Hinge Support    Secondary Buckle   "Rolling" Direction
306   Figure 4.  Early Development Unit Showing Natural Curvature , R2, of a Tape Spring  

Figure 5.  Tape Spring Element Geometry  

Figure 6.  Linear Shift Due to Hinge Stowed Radius Constraint  

Second Generation Design   With a goal of creating a more universal design, the hinge was redesigned to simplify the interfaces . The  redesigned hinge is shown deployed in Figure 7 and stowed in Figure 8. The fittings were redesigned to  mount to a flat panel rather than the tube frame . Because the hinge is open section,  the torsional stiffness  of the hinges is low and there is no need to have more than the two mounting screws since the hinge cannot  transfer any significant load in torsion.   R2  R1  α1  L  R2, Unconstrained   R2, Constrained   Linear Translation  resulting from  Constrained R2
307 The redesign better established and maintained the hinge curvature, mitigating the lateral motion  and  simplifying the release dynamics . Increasing the stowed radius also allow s use of a stiffer structure for the  mechanism  without the risk of overstress in the hinge springs . A tighter radius with sti ff hinge fitting  constraints could result in unacceptably high stresses  or reduced life. Increasing the stowed radius does  require more volume for the stowed hinge, but allows the  use of the hinge with a structure of any stiffness  without the risk of overs tressing the tape spring.   

The redesigned hinge incorporates shims between the hinge springs . This separates  the leaves of the  hinge to allow more room for the internal springs and eliminates the secondary buckle seen in Figure 3.  The direct electrical bond between adjacent leaves is be maintained with conduction through the shims ,  while providing clearance to eliminate the secondary buckling.    The acceptance test program was designed to allow substitution of a small test structure for the large flight   structure, allowing for a cost -effective test program using smaller facilities and increasing the available  facilities which can be used.    The resulting test program can be easily adapted to a variety of applications increasing the utility of the  design for reuse on future missions . With the experience gained from these projects, it is possible to build  and test affordable non- traditional mechanisms while maintaining traditional performance requirements.  

Figure 7.  Deployed Redesigned Hinge  

Figure 8.  Stowed Redesigned Hinge      Original Design Testing  

Original Design   To verify the performance of each hinge, output tor que testing was performed. The hinge fixture is shown  in Figure 9. The torque output is measured without constraining the translation of the rotating end of the  hinge. This is necessary because of two effects:  the center of rotation of the hinge changes as the end of