ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
52   Figure 13.  Actuator Development Hardware in Laboratory Test Stand 

Design of Twisting Tilt-Rotor Blade and Mechanism 

The Aerospace Corporation’s initial concepts of the rotor blade design involved a split spar and a semiflexible trailing edge that would let the blade twist by  shearing the two spar members. Since a robust  design could not close around the open trailing edge, an alternative was pursued. The team noticed that  shearing the two spar members also produced a sli ght separation between them. Using that knowledge,  the team pondered that the reverse methodology should also hold, i.e., separation of the members would  result in some amount of shear (and thus twist in  the blade). That supposition proved to be true and is the  principle that led to the design of the twist mechanism described below.   The mechanism design is based on a pin-follower, cam-based, detent shuttle concept. A CAD drawing of the mechanical system inside the blade was created, as shown in Figure 14 (some aspects not to scale). 

    Figure 14.  Preliminary CAD Representation of Twist Mechanism Inside Blade     End View0.4” End View0.4” NASA/CP-2010-216272
54 After the DADS model was created, several invest igative runs were performed to determine the  necessary force profile, specifically peak force and duration, to be delivered by the actuator. A series of  spike-and-decay force pulses with a 0.002 second fu ll-width-at-half-max (0.012 second overall) were  used, since that was the general pulse shape from t he actuator development testing, and the amplitude  was varied. According to the simulation, a 5.17 MPa (1.47 kN) [750 psi (331 lb)] peak pulse was not  successful in actuating the mechani sm past the detent, but a 6.89 MPa (1000 psi) pulse was successful.  With the latter pulse shape, the end-to-end actuation was predicted to be 0.004 s; a higher, 10.34 MPa  (1.97 kN) [1500 psi (442 lb)] pulse was predicted to actuate in 0.002 second.   So the actuator was tuned to the correct pressure  profile using appropriate levels of buffer gas and  electrolysis parameters, and the tilt mechanism hardware was built. A test fixture was also built at The Aerospace Corporation to simulate the DADS model (which in turn, simulated the rotor blade) with custom-designed stacks of Belleville washers at eac h of the three cam locations. Given additional  sources of fixture friction not accounted for in the DADS model, such as lug alignment rod sliding and  pillow blocks rubbing with the conne cting rods, the actual hardware an d test fixture actuation time of  0.007 second was within the expected range of uncertai nty. After the hardware was working satisfactorily  on that fixture, it was sent to Bell Helicopter for in tegration into their blade, as shown in Figure 16.    A total of thirty actuations were conducted at Bell dur ing integration bench testing, in order to properly  configure the actuator with the correct electrolys is time and current, buffer gas pressure, etc. An  additional thirteen actuations were conducted during the “run for record” (referred to as the Demonstration).  Three of the thirty bench test actuations ended on the up-stop (cams moving inboard in blade); and seven of the actuations ended on the down stop (cams moving outboard in blade); the remainder started or stopped somewhere in mid-stroke. There was only one bench test in which the actuator/mechanism  traveled the full stroke, from up-stop to down-stop, which resulted in a total of 4 degrees of twist . This  amount of twist met the goal of the project. High-s peed video was used to capture the twisting events,  with the beginning and ending frames showing total blade tip motion with respect to a protractor, as presented for one example in Figure 17.  

Figure 16.  Cams and Rods Installed In Blad e Spar Prior To Installation of Upper Skin  Once the team felt comfortable with the setup of the actuator and mechanism, the formal Demonstration for the DARPA customer was executed. Of the thirt een actuations conducted during the Demonstration,  all but the first one traveled the full stroke, stop-to-st op (the first was actuated from the mid-point, since  Guide Blocks  Aft Spar /  Lugs Fwd Spar  Actuator  Rod &  Push Rod  (not yet  connected)  NASA/CP-2010-216272
55 that is the installation configuratio n). However, those actuations only averaged about 2.5 degrees of blade  twist. 

Figure 17.  Beginning and End Frames from High Speed Video for Typical Sequence  After the Demonstration, the blade was X-rayed to determine if any internal failures occurred. The results,  shown in Figure 18 indicate that the rods in the most outboard cam may have been previously bent.  Some of the curvature seen in the X-rays is due to a “fisheye” effect because of projecting a point-source onto a flat receiver. However, the intersection of the inboard rod and Cam 1 does not appear to be  perpendicular.    

Figure 18.  Post-Demon stration X-Ray Results     Full Stroke 100% to 0% -4.5o -1.5o Full Stroke100% to 0% -4.5 o -1.5o Cam 1 – Up Position (Inboard)Cam 1 – Down Position  (Outboard) Cam 2 – Up Position  (Inboard) Cam 3 – Up Position  (Inboard)Cam 2 – Down Position  (Outboard) Cam 3 – Down Position  (Outboard) Cam 1 – Up Position (Inboard)Cam 1 – Down Position  (Outboard) Cam 2 – Up Position  (Inboard) Cam 3 – Up Position  (Inboard)Cam 2 – Down Position  (Outboard) Cam 3 – Down Position  (Outboard) NASA/CP-2010-216272
56 Post-Demonstration Evaluation 

The actuator proved to have more than sufficient force to actuate the mechanism assembly. The loads in  the Demonstration tended to be a little over 2.2-kN ( 500-lb) tension for moving the cams inboard in the  blade and around 1.78-kN (400-lb) compression for movi ng the cams outboard. These forces are roughly  what the DADS model predicted (1.5-2.0 kN (330-440 lb) minimum for a successful actuation), but are  slightly higher in reality due to the additional  friction features in the real hardware.   When the model used lumped masses at each of the lug locations to simulate the mass of the twisting blade, and used a pulse of 6.89 MPa (1000 psi) over 2 ms (full-width, half-max) it predicted a profile  shown by the red curve in Figure 19, for actuator push rod load vs. time. That analytical curve is overlaid  onto one of the successful bench test sequences for comparison.  Figure 19.  DADS Output (Red) Overlaid  Onto Bell Test Data for Sequence 30     Lessons Learned 

1. Electrolysis/combustion actuator s can be used successfully to achi eve a higher specific power than  other currently available technologies.  2. Electrolysis time and buffer gas volume can be used to effectively tune the combustion pulse in both  peak magnitude and time duration to meet specific application needs.  3. Intelligent solutions can sometimes unexpectedly re sult from lofty goals. The actuator goal of 16.7  kW/kg (50 hp/lb) was chosen because it was 1- 2 orders of magnitude above what had ever been  produced previously. Initially thought by the research team to be unattainable, that “stretch” goal drove the team to produce innovative hardware.  4. Non-metallic materials should always be part of the design trade space. Some spacecraft and launch  vehicle engineers can be lulled into the familiarity of metals  and their properties , but lessons can be  learned from the air vehicle industry which might be just as concerned or more about light weight  solutions. The pillow blocks, made out of a sintered polyamide, we re a wise choice for strength,  weight, and low friction.  5. Befriend the machine shop experts. Present them  with reasonable designs and more importantly,  reasonable deadlines. Similarly, solicit their input  during the design phase, not after it has been  frozen. Chances are, they have seen many different  design solutions and may have different ideas to  offer. 

   NASA/CP-2010-216272
57 Summary/Conclusion 

Extremely powerful, reusable actuators can be co nstructed by generating hydrogen and oxygen through  electrolysis and then igniting the gas mixture when actuation is desired. The amount and rate of gas  generated is simply a function of electrolysis curre nt and time; the more Coulombs delivered, the more  gas will be generated, and the more power will be available.  This technology was applied beyond the theoretical env ironment to achieve both goals for which the team  set out to achieve. A test fixture was created to meas ure the instantaneous specific power of the device,  which was measured at over 46.8 kW/kg (140 hp/lb), or three times the proposed goal. This result  demonstrated that this type of actuator is outside the current state of actuator performance in terms of  power density. Furthermore, a practical application of using such an actuator to twist a prototype rotor blade was identified, and a development actuator and power transfer mechanism were built. The goal to  achieve 4 degrees of blade twist was realized with the development hardware.   Acknowledgments 

The authors would like to thank the Defense Advanced Research Project Agency for their guidance and  support of this research, which was part of a proj ect led by Dr. Louis Centolanza entitled “Controllable  Active Materials Via Internally Generated Pressu re,” under U.S. Army Research, Development, and  Engineering Command (RDECom) contract W911W6-05- C-0015. In addition, Dr. Michael J. O’Brien  (materials), Aldrich R. De La Cruz (FEM modeling), and Dr. Brian M. Gable (materials and heat treating)  were all instrumental in helping the project. The authors also wish to acknowledge the excellent assistance of Nolan Phillips, Ernie Powell, Joe Tzeng, and the entire Bell Helicopter partner team.        NASA/CP-2010-216272
61 MAPLUB SH050-a Grease NYE 2001A  MAPLUB SH100-b Grease NYE 2001A  MAPLUB SH051-a Grease NYE 2001A  MAPLUB SH101-b Grease NYE 2001A  Braycote 601EF Grease Castrol Brayco 815Z  Braycote 601EF Micronic Grease Castrol Brayco 815Z  Rheolube 2000 Grease NYE 2001  Sputtered MoS 2 Solid --  Sputtered Lead Solid --     Apparatus 

Spiral Orbit Tribometer   The Spiral Orbit Tribometer is a test facility developed by NASA Gl enn Research Center and recently  purchased by ESTL to advance the assessment of l ubricants and coatings for space applications. The  facility reproduces the kinematics of an angular contac t bearing, and allows for the evaluation of friction  and degradation rates (i.e. consumption) of lubricants in detail.   The Spiral Orbit Tribometer is essentially a thrust bearing, with a single ball held between two  interchangeable flat plates, located within a vacuum chamber. A load is applied to the top plate via a spring-loaded linear translator. The lower plate rotates via a motor located outside the chamber, causing the ball to move in a spiral path with a radius ~21 mm. This configuration causes the ball to spiral outwards, and a fixed guide plate is positioned to keep the ball within the flat plates and to maintain a  repeatable orbit. The region of each orbit for which the ball is in contact with the guide plate is denoted as the scrub (see Figure 1). A force transducer behind th e guide plate measures the force exerted by the ball  onto the guide plate. From this the friction coefficient fo r each orbit is found, and can be plotted to give the  performance of the lubricant over time (Figure 2). 

Figure 1. Internal arrangement of SOT, showing  flat and guide plates  Figure 2. Typical friction plot created with SOT  (using grease MAPLUB PF100-b) showing  typical regions sampled for friction coefficients 

The arrangement of the SOT allows the ball to experience rolling, sliding and pivoting – motions experienced by a ball in an angular contact bearing. This  allows for a more representative testing of a  lubricant than conventional pin-on-disc test ing, which only recreates sliding motion.   Samples

Test plates were supplied along with the SOT by Sp iralab LLC, Cleveland, OH. All samples (flat and  guide plates) were manufactured from 440C stainles s steel, and polished to a surface roughness   100 values  around 400th  orbit 100 values around test  midpoint  NASA/CP-2010-216272
62 Ra <0.05 microns. Balls used were either 12.7 mm (1/2 inch) or 7.14 mm (9/32 inch) diameter,  manufactured of 440C and 52100 steel respectively, depending on the requirements of the particular test,  explained in more detail below.  Controllers

The SOT is controlled by a supplied laptop PC, running a Labview-based data acquisition program. 

Procedures 

Sample preparation   Due to the limited number of test specimen sets, it proved necessary to re-prepare used samples for subsequent testing. This was performed by the National Centre of Tribology. Samples were polished using a diamond paste solution mixed with Meta-Di, a water-based diamond suspension solvent, on a 6-micron polishing disc. Samples were cleaned before and after polishing in methanol, and dried using a hot air blower with dry cotton wool. Balls were not reused for multiple tests.  Prior to testing all balls and plates were solvent  cleaned in a Kerry cleaning plant using Lenium ES  solvent in accordance with standard ESTL practice.  Lubrication

Lubrication was applied to the balls only. Liquid lubr ication was achieved through the preparation of a  solution of lubricant diluted in an appropriate solvent, of a known concentration. This solution was applied  directly to a rotating ball, and the solvent allowed to evaporate from the ball’s surface. The amount of  lubricant was determined using a high accura cy 6-point microbalance (accurate to 1 μg). A typical  lubricant amount of 50 μg was applied for each test. 

For grease lubrication, the ball was weighed dry and a minimal amount of grease (less than 1 mm3)  applied directly to the surface. The ball was then rolled successively between three pairs of solvent  cleaned Ultra Clean Level 100 Polyethylene tubing, stretched over Petri dishes, to evenly distribute the grease (Figure 4). The ball was subsequently re-w eighed, with the weight change being the grease  uptake. A typical lubricant amount of 50 μg was applied for each test. 

Figure 3. Liquid lubrication – application of oilsolvent solution directly to the ball  Figure 4. Grease lubrication – grease  lubrication through rolling between  polyethylene sheets    Solid lubrication was performed with ESTL’s sputter coating rig, coating the balls only, to a desired  thickness. The coating thicknesses were assess ed using a calibrated X-Ray Fluorescence (XRF)  measurement system, taking 20 measurements of 60 seconds for each coating run. Coating thicknesses for each test are detailed below.     NASA/CP-2010-216272
63 Test Program 

Testing was performed using the SOT under vacuum (<1.3 x 10-4 Pa, 10-6 torr) at room temperature  (~23deg.C). Tests ran until the friction coefficient exceeded 0.3 for three consecutive orbits, and which  point the motion was halted by an automatic trigger. Other test details are given in the relevant sections  below.  Liquid lubricant assessment

Tests on each liquid lubricant were performed over a range of mean contact stresses (1.00 – 1.75GPa). For all but the highest contact stress, a 440C ball of 12.7-mm diameter was used (a smaller 7.14-mm ball of 52100 steel was used to achieve the higher contact stresses without exceeding the limit of the linear  translator). Tests on the Fomblin oils were perfor med at 30 RPM, and increased to 100 RPM for the NYE  oils due to the expected longer lifetimes of these oils. Preliminary testing at ESTL using Z25 demonstrated no apparent dependence of lifetime upon ball size or rotation speed for this lubricant within this range.  Grease lubricant assessment

Grease lubricant tests were performed using a 12. 7-mm diameter 440C steel ball, rotated at 100 RPM  and loaded to 1.50 GPa mean contact stress. A minimum of three tests were performed for each grease formulation, and results presented in Table 2 show the mean values.  Solid lubricant assessment

For all solid lubricant tests, a 52100 steel ball of 7.14 mm diameter and rotation speed of 100 RPM was used. For MoS 2, the coating thickness was varied with constant mean contact stress. For lead, the mean  contact stress was varied with constant coating thickness.   Table 2. Test matrix for solid lubricant tests   Lubricant (thin  solid film) Coating thickness  (Angstroms) Mean contact  stress (GPa)  MoS 2 800 1.50  MoS 2 2300 1.50  MoS 2 5300 1.50  Lead 850 1.50  Lead 850 1.75  Lead 850 2.00 

Liquid Lubricant Assessment 

The frictional behavior of all oils tested was broadly si milar. Start-up friction values were typically μ = 0.1.  Steady state friction is then mainta ined until rapidly increasing to above μ = 0.3 (see example Figure 2).  This increase in friction is accompanied by a slight increase in chamber pressure, indicating the presence of volatile constituents – an expe cted observation as the lubricant deg rades [4]. Figures 5 & 6 show the  lifetime and friction coefficients of the test oils as  a function of mean contact stress. Lifetimes are  normalized to orbits/microgram of lubricant. Steady stat e friction coefficients are calculated by averaging  100 readings around the 400 th orbit of each test, to allow for comparison with previous results.  NASA/CP-2010-216272
64 Figure 5. Lifetimes of liquid lubricants as a  function of contact stress Figure 6. Friction coefficients (assessed at 400th  orbit) of liquid lubricants as a function of  contact stress 

Lifetimes were found to be much reduced for the PFPE oils (Fomblin Z25 & Z60) in comparison to the MAC oils NYE 2001 & 2001A, a result previously s een for these oils under boundary lubrication  conditions [5], as well as from past experiences with bearing tests at ESTL, operating in both boundary and mixed lubrication modes [6]. There is also a small but  clear distinction in the friction coefficients of the  two groups. A comparison of this data with prev ious studies performed using a SOT by NASA on the  lubricants Z25 and 2001A [7, 8] reveals good correlat ion between the friction and lifetimes of 2001A, but  poor correlation when considering the lifetimes of Z25 oil. Lifetimes of Z25 measured by ESTL are some 3x greater than those found in [8]. The likely cause of this discrepancy lies in the differing sample  cleaning techniques between the two studies, as surface co ndition is thought to play a significant role in  the degradation rates of PFPEs [9].   The performance of Z60 was very similar to that of Z25, with slightly reduc ed lifetimes and comparable  friction coefficients. Similar results have been observed for these Fomblin oils in non-boundary lubricating  conditions [2]. No significant difference was observed between the two NYE oils under the test conditions.  This result is curious when we consider that the oil NYE 2001 contains additives to improve this lubricant’s boundary performance. Ho wever, it is proposed that due to the minimal amounts of oil tested  within the SOT (typically 50 µg), re sulting in only a few atomic layers of lubricant on the ball’s surface,  these additives are not replenished once they are remov ed from the contact zone. Typically there is an oil  reservoir present to replace the additives.   All oils displayed a decrease in lifetime, and a correspo nding slight increase in friction coefficient, with  increasing contact stress, an effect similarly seen in previous studies using a SOT [7]. This behavior  demonstrates the sensitivity of these lubricants to contact stress when operating under boundary  lubrication conditions. The cause of the potentially anomalous results seen for the MAC oils running at  1.75 GPa is not clear.  Post-test inspection of the samples showed markings on the flat and guide plates from the running of the  ball. Inspection with a low powered optical microscope re vealed these marks to consist of brown material  deposited away from the ball tracks; consistent with the residue of the consumed lubricant.  

Grease Lubricant Assessment 

Figures 7 & 8 show respectively the lifetime and fric tion coefficients of the assessed greases. Lifetimes  are normalized to orbits/microgram of lubricant, an d friction coefficients are taken as the mean of 100  readings around the 400 th orbit, and the midpoint of each test (Figure 2). Mean contact stress for all  grease tests was 1.50 GPa.   NASA/CP-2010-216272
65   Figure 7. Lifetimes of various greases assessed under vacuum at 1.50 GPa mean contact stress

Figure 8. Friction coefficients (assessed at 400th orbit and mid-test) of various greases assessed  under vacuum at 1.50 GPa mean contact stress

In general, friction profiles generated from grease lubric ation were similar to those of the base oils, with  long periods of low friction before a comparatively r apid increase to failure. It was observed that the  increase of friction coefficient during the test was mo re progressive for the MAC greases in comparison to  NASA/CP-2010-216272
66 the PFPE-based greases. This effect has been observed previously when testing greases using a SOT  [10, 11].  All tested PFPE-based greases showed little variation in lifetime, with a value of ~200 orbits/µg for all  greases (Figure 7). This value is significantly in creased in comparison to studies performed at NASA  under similar conditions (30-40 orbits/µg) [11, 12], and as stated above, this is believed to be a consequence of the differing cleaning techniques implemented. Considering lifetime the inclusion of MoS 2  filler was not found to cause a significant effect, nor was a difference observed between the old (-a) and new (-b) MAPLUB formulations, nor inde ed between Braycote 601EF and 601EF Micronic.   Considering the friction coefficients (Figure 8) we find the new MAPLUB PFPE greases display lower friction coefficients than their comparative old formulat ion greases. This difference is most exaggerated in  the early stages of the test. When plotting friction coe fficient against orbits, it is seen that the new (-b)  greases undergo a change in gradient after a few thousand revolutions (Figure 9), not seen for the old  (-a) greases (Figure 10). This difference in the frictional behavior of the greases is the cause of the lower friction coefficients for the new PFPE greases, and is seen when considering greases both with and without MoS 2 filler. The friction coefficients of the Bray cote greases fall somewhere between the old and  new MAPLUP formulations.   Lifetimes for MAC greases were greatly extended in comparison to the PFPE greases, again in agreement with bearing grease tests at ESTL. Lifetime s for new formulation MAPLUB greases (-b) were  seen to be reduced by an order of magnitude in comparis on to the old (-a). It is thought this difference  can be attributed to the greater viscosities of t he new formulations, hampering the re-introduction of  grease into the contact surfaces during the test s. In addition, lifetimes for greases containing MoS 2  thickener were found to be less than those containing only PTFE (e.g., SH101/100 and SH050/051). This result is interesting when we consider that the MoS 2 content of these greases is ~1% by volume. Given  the minimal amount of grease used per test (~50 µg) it  is somewhat surprising to see the addition of such  a small amount of MoS 2 apparently effecting the lifetimes to such a degree.  

Figure 9. Early stages of PH101-b test,  displaying change in gradient around 2,500  orbits Figure 10. Early stages of PF101-a test, no  gradient change   Rheolube 2000 was found to have similar friction to  that of the MAC-based MAPLUB greases. This  finding is akin to that found in [13], where the performance of low-speed bearings showed comparable  torque for Rheolube 2000 and MAPLUB SH050-a & SH051-a. Bearing tests at ESTL have also  demonstrated lower torque for Rheolube than Braycote 601 when rotating at low speeds [14], again in accordance with the results generated with the SOT.  Inspection of the samples post-test revealed brow n deposits away from the ball track, similar in  appearance to those of the base oils, consistent with the degraded lubricants.  NASA/CP-2010-216272