ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
464 occurred. A post -test functional verification showed that the actuator remained fully operational after the  test. It proves that the  design is able to survive the required l aunch vibration loads . 

Synthesis  of Actuator Characteristics  

Table 2.  Actuator main  characteristics  

Characteristic  Value  Requirements   Mass :  118 g   < 200g   Overall dimensions :  89 x 44 x 44 mm3 100 x 60 x 60 mm3  Lifetime : 1055   (stopped, no damage)  > 1000 cycles   SMA wire lifetime : > 7657 cycles  --    49.8 W (27.8 V, 1.79A)  < 50W (28V)   Verified output force:  96N* 165N  Latching position:  1.5 mm  < 1.7mm   Maximal stroke:  2.5 mm  > 2.2mm   Push duration:  6.5± 0.4 s 5s < t < 60s   Pull duration:  11.9± 5.4 s 5s < t < 60s   Max operating temperature:  120°C  > 70°C   Launch vibration load:  verified  28.3 gRMS  *The quality of the produced CuAlX monocrystalline wire samples was not sufficient to reach the cycling  stability under  the required force. It was therefore decided to reduce the test load at 96 N (from 16 5 N) for  the test campaign . Recently , industrial quality level CuAlX wires are commercially available and should   meet the design load of 17 5 N. 

Conclusion  

A bi-stable actuator based on a high transition temperatures Shape Memory Alloy  has been successfully  designed, manufactured and tested. This mechanism, able to provide slow -acting pushing and pulling  forces , showed great repeatability over a high number of actuating cycles demonstrating high overall  reliability  of the selected design .    One major challenge was to design a SMA heating system compatible with the use of high transition temperature SMAs that are highly conductive and have shape- memory strains up to 10%. A novel heating  system has been developed and successfully tested on single wire setups as well as on a fully integrated  mechanism.     Through this project, the technology based on high transition temperatures SMA reached the readiness  level TRL -4. Further material characterization tests on industrial quality CuAlX wire as well as additional   environmental tests are to be performed to reach TRL -5.
466 Friction Holding Brakes  

We have developed a disk type of friction brake for a rotating antenna drive application on a drone, not  directly for space but for stratospheric use. It has  therefore been qualified to environmental conditions, very  similar to space –  low air pressure (100  mBar) and low temperatures ( -60°C).  

In order to achieve high torque,  the design makes use of the lar ge diameter of the torque motor, which is  to be secured by the brake when  the motor is  stationary or inactive. A thin stainless steel ring is mounted  around the rotational axis, attached to the rotor of the motor.  This ring has high torsional stiffness but is  flexible in the axial direction  to allow for some freedom of motion of the mechanism in the shaft axis .   Two brake calipers with pinching brake pads  are mounted around the disk. These calipers are operated by  mechanical springs  to apply braking friction force to the disk and thereby braking torque to the motor .  Electric solenoid actuators release the pads and reduce the brake torque to zero.  

Figure 1.  10-Nm Friction Disk Brake (MACCON)   This brake also fulfills  another demanding requirement. Due to the sensiti vity of the application, it was  specified that a specified maximum level of reaction torque on the supporting structure may not be  exceeded  if the brake is operated,  while the motor is rotating.  The brake described requires  36 W to release  but only 7  W to maintain the released condition. Due to the springs incorporated in the c alipers this brake  is fail- safe.    Now we describe another type of friction brake, which can achieve high braking torque values in a much  small er diameter  thanks  to the large  torque multiplication factor of  a gearbox . Figure 2  shows a high torque  brushless DC actuator with an integral gearb ox and f riction disk brake. This actuator is used in a robotic  arm application where power -off holdin g of a large moment is required. While the peak torque of the  brushless actuator is 22 Nm, the unpowered holding torque of the brake is 8 Nm  reflected to the output  shaft . With the mechanical advantage of the gearbox, a small  friction disk brake provides  significant power - off holding torque provision  for a minimum of mass and volume.    A disadvantage of both friction disk brakes  is the relative  high power needed to pull -in or activate the freerunning condition  of the actuator . For instance, the brake design in the actuator in Figure 2 draws  7 W  power at 28 volts DC, the rated voltage. The voltage and power to maintain free- running actuation is
467 however significantly lower, as it takes less than one watt holding power to maintain the brake in the  disengaged condition.  

There are several methods to minimize the sustained power draw of both friction brake designs  described  above. Sin ce the “drop- out” or engagement voltage is significantly less than the pull -in voltage, the control  system may simply reduce the excitation  voltage after  the free-running condition has occurred. The  activation happens qu ickly, in about  10 ms.  As it may not be desirable to regulate the voltage, it is simpler  to have two separate coils in the brake bobbin assembly . This is also an option for redundancy . The second  “low power” or “sustaining” winding may be  used to allow d eactivat ion of the first “high power” or “activation”   winding after the free- running has been  achieved. The “sustaining” coil winding may have significantly  higher resistance to minimize power draw  compared to the “activation” coil winding. This schematic is  represented in Figure 3.  

Figure 2.  22- Nm Actuator with Integral Friction Disk Brake (A vior) 

Figure 3.  Dual Coil Brake Schematic (Avior)
468 A method to easily remove the power from the “ activation”  coil is to add a “blocking” capacitor in series with  the winding.  As the DC excitation is applied to the “ activation ” coil winding, the current flow s through the  winding, as the inductance allows . Once activated,  the current levels off and the blocking c apacitor drains  the current dow n to zero . This is a simple method that does not require additional logic or switches to  disable  the “activation” coil .  

Passive Reluctance Brakes  

Passive reluctance brakes (or d etent brakes) are always active, which in turn means that the motor needs  sufficient starting torque to overcome the brake detent torque in addition to any stall load torque. However ,  once the shaft is rotating, the additional power needed to run the motor drops, as the reluctance torque equally assists as well as resists motion; the effect of this pulsing load torque drops with increasing speed.   Typically this type of brake only offer s a limited number of holding positions per revolution ( 4,6,8,12 etc. ).  

Figure 4.  Stepper Motor and Reluctance Brake Torque vs. Position Plot (Avior)    Figure 4 shows a Motor Detent Brake Torque versus position (or Torque- Theta) plot  of a stepper motor  combined with a passive reluctance brake. In this configuration, the motor steps and stable brake detent  positions coincide. At each step of the stepper motor, the torque drops off from the peak torque to zero in  a cosine function (for sinusoidal back EMF ).  The detent brake repels against the mot or torque for the first  half of the step, and assists in the second half of the step. While the area under the curve integrates out, it is over -simplistic to state that the detent brake does not have an impact on dynamic performance. In fact,  the detent brake may have a positive  affect on performance. S ince it can add torque as the stepper motor  is approaching the stable step position, the rotor accelerates faster through zero and creates more overshoot at cardinal stepping frequencies.    As described in [1] , overshoot in a stepper motor creates cardinal maxima and minima torque variations at  various points along the torque- speed curve. The addition of the detent brake inherently attenuates the pull - in torque of stepper motor actuator at low pulse- rates . The actuator pull -in torque curve in Figure 5 shows   performance with and without a 19- mm passive detent brake that produces 1.0- Nm torque at the output .  As you see at low velocities, the whole torque of the detent brake reduces the available pull in torque. As  the step rates increase, there are varying affects of the detent brake. The pull -in torque attenuates or
469 increases at varying step rates . At high velocities, above the mechanical time constant, the detent brake  has virtually  no affect on the actuator perf ormance.  

The advantages of the passive brake configuration are zero energizing power and no control circuit  requirement.  

Using this brake directly on the servomotor shaft before a high- ratio reduction gearbox allows its dimensions  to remain small and mi nimizes servocontrol problems, which result from the motor having to operate against  the high ripple torque of the reluctance brake. Positioning is intentionally limited to the preferential positions of low reluctance to which the brake is naturally biased.  It is interesting to note that the pull -out torque is  much less affected by the d etent brake. In other words, the pull -out torque performance of an actuator with  a detent brake closely mat ches the pull -out performance of the same actuator without a d etent brake. We  strongly recommend full characterization with simulated load inertia of actuators incorporating passive detent brakes.  

Figure 5.  Pull-In Torque Performance      Active Reluctance Brakes  

This active reluctance type of brake will be used in a valve actuator of a space vehicle launcher. The solution  chosen has a special stator/rotor geometry with  many positions of minimum reluctance per revolution.  Permanent magnets fitted in the stator  cause this brake to have 200 preferred holding positions (of 1.8°  pitch) with a closely defined holding torque value.  The permanent magnet field can however be neutralized  by passing current through a single winding in the stator, thus allowing the mechanism to rotate with  minimum magnetic  drag.  Energizing the coils causes the flux coupling between stator and rotor to be  shifted  into the stator back iron resulting in a magnetic short -circuit  for the permanent ma gnets.   

The first application we have realized is illustrated in Figure 6. It has an outer diameter of 120 mm, a length  of 30 mm, and a holding torque of 3.3 Nm. The torque characteristic of the active reluctance brake over  one pitch is shown in Figure 7 and the flux distribution in both operating conditions in Figure 8.
471 *   Figure 8 B.  Magnetic Flux distribution in active rel uctance Brake with stator coils not energized  (MACCON)   The clear advantage of this brake design is the high number of preferred braking positions per revolution,  200, without the need to energize the brake. There is also a good geometrical match between the brake  and its matching motor design dimensions . On the negative side it does require high excitation power   (58 W) to release, i.e. , for the neutralization of its holding torque.  However i n applications  where release is  only needed for a short periods , this energy dissipation can be accepted.  

Conclusions  

The four holding brake configurations  described in this paper have been especially designed for demanding  high reliability a nd space applications. As they offer very differing mechanical configurations and braking  characteristics, it is probable that one of these configurations will serve any new electro- mechanical dri ve  or actuator application.    All brake designs discussed share the essential properties of requiring zer o energy, when braking torque  is required and of being fail -safe. If the energy supply fails, the holding condition is guaranteed.  Passive - reluctance brakes offer the advantage of no power excitation whatsoever , but detailed characterization of  performance with simulated load inertia is recommended, due to induced operational pull -in torque  variations and thus an increased step- response overshoot . Additionally, passive reluctance brakes are not  recommended for use with servo -systems, because brake torque may be a significant source of load  disturbance for the servo control system .    Lastly, all the active brake configurations can be made electrically redundant , with the simple  implementation of a second excitation or release c oil [2] . Friction disk brakes designs may use an   unbalanced power configuration with a two- coil option  to minimize quiesc ent power draw by means of  a  higher resistance sustaining coil.  

References    1. Scott Starin, Cutter Shea, “Primer – Stepper Motor Nomenclature, Definition, Performance and  Recommended Test Methods”, Proceedings of the 42 nd Aerospace Mechanism Symposium,  NASA  Goddard Space Flight Center, May 14- 16, 2014   2. Ted Hopper, Dr. Markus Anders, Christoph Stuckmann,  "Building electric motors for space, with  redundancy and high reliability ”, Proceedings  of the 14 th European Space Mechanisms  and Tribology  Symposium  2011
474 for consulting commercial customers to get feedback on their future needs and requirements neither  the  commercial nor the institutional customers were aware of the development.  As a consequence a very  efficient and rapid development could be realized.  

Project D efinition  

The project definition described here presents the approach implemented in the Future Actuator  development  program. The entire development program is divided in three phases:   1. Requirements definition and concept phase (without customer)   2. Concept validation and performance review (without customer)   3. Concept finalization ( together  with customer)   The project budget was small and therefore the program had to be planned carefully. For this reason,  the  Research & Development  included only phase 1 and 2 . The third phase was postponed until the test results  were available. In the meantim e the third phase – has been initiated an d is running.  

Market Survey   Before the requirements and concept definition phase , a market survey was performed. The market survey  showed that there are a number of products on the market with very similar performance to the RUAG  standard actuator SARA21 . Additionally , there is a tendency towards actuators with higher gear ratios  and  small er step sizes  as well as extreme price pressures from the customer . 

Requirements Definition  The requirements definition was split in two areas, the programmatic and the technical requirements . For  the programmatic requirements,  there was the question where the actuator sh ould be positioned on the  market. There were three possibilities:   1. Develop a low price  product with similar or reduced performance (compared to the SARA21)   2. Develop a product with improved performance and same price  as SARA21   3. Develop a high end product with higher price   Based on the feedback from key customers , it was known that next to per formance the price is the most  important factor for the product. Therefore, RUAG decided to focus on a combination of performance  improvement in parallel with a price reduction.           Figure 1.  RUAG’s SARA21
479 were investigated. On the one side there were different means how the magnetic field can be generated  and on the other side the sensor c ould be combined with other  sensor types  to achieve the r equired  accuracy . Many solutions were withdrawn for cost, space or other r easons . The solution finally selected for  the Future Actuator is simple and flexible. The sensor head itself is very compact and shows  a high  resistance against radiation  and vacuum  as well as a long lifetime.      The sensor offers its potential only in combination with the electronics. Different  commercial sensors with  integrated electronics  are available. However, terrestrial  electronics are in most cases not compatible with  space requirements . A qualification of a commercial non- space electronics was not considered as the  qualification risk of such a device is too high. In addition, it is preferable to integrate the electronics in one  box instead of having each actuator equipped w ith its own radiation -hardened electronics.  Therefore, the  focus was on the sensor itself and the development of an algorithm to achieve the needed absolute position  accuracy.  

Breadboard Definition   The Future Actuator  design phase was focused on the flight design but for the breadboard model  compromises were needed to comply with the small budget and the schedule constraints. Therefore,  the  design w as modified for the breadboard. With this approach a good mix between a cheap and fast  breadboard model together with high representativeness was achieved.  

          Figure 5.  Future Actuator  Breadboard Model  

One important modification was the removal of some coatings. Based on RUAG experience , the removed  coatings mainly  had an influence on the lifetime and the lifetime was not tested  on the breadboard.  The  other modific ation was the motor interface.  As the final motor  should be selected based on the test results  the motor could not be selected at this stage. Therefore,  the motor interface was kept modular to comply  with the test environment.  

Collaboration with Harmo nic Drive  AG   Gears based on the strain wave gear principle have a remarkable record in space applications  [1]. They  are backlash free and enable high ratios  while being compact. This results in high torque density and low  weight solutions. Furthermore,  strain wave gears enable high accuracy and reversibility of motion. All these  features/ characteristics were required. Harmonic Drive AG, Limburg, Germany, proved to be a capable,  strong partner in previous  application related  activities and targeted development of space rated Harmonic  Drive® gears based on the strain wave gear principle.  Based on a basic set of technical requirements,  collaboration with Harmonic Drive AG was initiated in an early project phase.  There was no frozen  specification which had to be fulfilled . On that basis a fruitful, open minded coo perative work environment  was set . This was very appreciated by both parties, as  it enabled elaborat ion of the best solutions  using  knowledge and resources from both sides, RUAG and Harmonic Drive AG.
481   Figure 6.  Double -Stage Gear after assembly at Harmonic Drive AG  

The trend of the transmission accuracy of one gearbox  during operation is shown in Figure 7. The  transmission accuracy of the gear represents a linearity error between input and output angle and is defined as the sum of the maximum positive and negat ive (see “min/max” in Figure 7) differences between  theoretical and actual rotation angle. As expected, analyses showed that transmission accuracy of the 2nd stage is dominant. The transmission accuracy of the 1st stage has a minor influence. The measured values  of the transmission accuracy of all manufactured component sets are shown in Table 1. All component sets  showed good values which were within the specified range (<2.9x 10 -4 rad). Beside the transmission  accuracy , the torsional stiffness of each component set was measured. Torsional stiffness is calculated  from the measured tor sion angle when applying a load at the output shaft whilst mechanically locking the  input shaft. The torsion curves are shown in  Figure 8. Calculated stiffness v alues are presented in Table 2. 

Figure 7.  Transmission Accuracy    Figure 8.  Torque – Torsion curve  

Table 1.  Transmission accuracy   Component Set –  No. Direction of  rotation  Transmission Accuracy   *10-4 [rad] [arcsec]   001 cw 1.2120 25  ccw 1.2605  26  002 cw 0.8727  18  ccw 0.8242  17  003 cw 1.4060  29  ccw 1.5514  32
482 Table 2.  Measured Torsional S tiffness   Component Set –   No. Stiffness   K1 [Nm/rad]  Stiffness   K2 [Nm/rad]  Stiffness   K3 [Nm/rad]   0…14 Nm  14…48 Nm  48…67 Nm   001 40500  42400  46300   002 28700  41000  45400   003 31500  42400  46200  

Testing at RUAG  

The Future Actuator test program had to mitigate all identified potential risks and to ensure the mechanism  was properly understood before going into Phase 3 of the development . Therefore, the Future Actuator test  program had two main objectives.   1. To characterize the performance of the ac tuator without the motor   2. To derive the needed inputs for the motor sel ection   Based on this , the test program was tailored to cover the above defined needs and a comprehensive test  program was defined in order to properly characterize the Future Actuator breadboard model. The following breadboard tests were performed:  

Run-InNo load friction  test @ ambientFriction test under  load @ ambient  & TVStart torque   measurement Bending  stiffness testTorsional s tiffness  testAxial and  lateral s tiffness  test  Figure 9.  Future Actuator breadboard test program  

A life test was not included in the breadboard test program as the lifetime requirements for the actuator are  not very demanding. For the motor , the lifetime will be more challenging due to the high gear ratio but the  motor will be subject of Phase 3. The fact that the Future Actuator has exactly the same interface to the  spacecraft and the similar overall dimensions to the SARA21 allowed to use existing test infrastructure.  This circumstance saved the project a lot of time and effort . Therefore the preparation time for the tests  was considerably small.  

Run In Test and No Load Friction Test   It is known that the strain wave gear s have a run in effect  [1] and therefore it was important to under stand  this effect with a double- stage strain wave gear. For this test the no load t est setup  was used (see  Figure  10). As the first gear stage is the dominant one for the friction torque,  also the first stage was dominant  for  the run- in effect.    The Future Actuator run -in test results showed that the torque reduction and behavior during the run -in of  the double- stage Harmonic Drive® g ear was very similar to a normal single- stage Harmonic Drive® g ear.  The torque reduction is in the range of 15%.    In the middle of the run- in the rotational  direction was changed which can be seen in Figure 10 as a tor que  increase for a short sequence in the middle of the test sequence. This behavior is in- line with standard  strain wave gears. After the run- in, the no load running input torque was in the range of 6 to 10 N -mm  depending on speed and start condition.