Cryocoolers 11

CRYOCOOLERS 11 A publication of the International Cryocooler Conference CRYOCOOLERS 11 Edited by R. G. Ross, Jr. Je...

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CRYOCOOLERS 11

A publication of the International Cryocooler Conference

CRYOCOOLERS 11

Edited by

R. G. Ross, Jr. Jet Propulsion Laboratory California Institute of Technology Pasadena, California

KLUWER ACADEMIC PUBLISHERS NEW YORK, BOSTON, DORDRECHT, LONDON, MOSCOW

H%RRN ,6%1 3ULQW,6%1

0-306-47112-4 0-306-46567-1

.OXZHU$FDGHPLF3XEOLVKHUV 1HZ 1.1 watt of cooling at 80 K with a compressor specific power of less than 25 watts/watt. An analysis of available tactical cooler compressors led to the selection of an advanced 1.75 W tactical Stirling cryocooler manufactured by DRS Infrared Technologies (formerly Texas Instruments). The particular model, shown in Fig. 1, is based on an advanced linear compressor with its two pistons operated head-to-head and supported on flexure springs to achieve long life and good vibration suppression. This new flexure-supported compressor is one of a family of advanced flexure-equipped compressors being developed to achieve extended-life tactical coolers.1

Figure 1. DRS 1.75 W tactical cooler with drive electronics.

GAMMA-RAY PT COOLER DEVELOPMENT AND TESTING

157

Figure 2. Refrigeration performance of the DRS 1.75 W Stirling cooler as a function of input drive voltage, coldtip temperature, and coldtip load.

The thermal performance of this cooler, using the Stirling expander shipped with the cooler, is shown in Fig. 2. The more conventional DRS 1.75 W cooler has similar performance, but uses helical coil springs to support the pistons and has a predicted life greater than 5000 hours. JPL has had good success using the conventional (non-flexure-spring) DRS 0.2-watt, 1-watt, and 1.75watt Stirling coolers for a variety of low-cost, intermediate-life space missions.2,3,4 PULSE TUBE DESIGN AND CONSTRUCTION

The second task critical to achieving the required cooler performance was the development of a high efficiency pulse tube expander carefully matched to the compression attributes of the DRS

compressor and the interface requirements of the JPL gamma-ray detector mounting system, shown schematically in Fig. 3. This task, carried out by Lockheed Martin Advanced Technology Center, involved first thoroughly characterizing the DRS compressor, then designing and fabricating a pulse tube consistent with the compressor and the gamma-ray detector cooling load and mounting interfaces. The chosen concept was the U-shaped pulse tube shown in Fig. 3.

158

SPACE PULSE TUBE CRYOCOOLER DEVELOPMENTS

Compressor Characterization To achieve an efficient design for the proposed pulse tube it was necessary to include an accurate model of the compressor's performance in the overall pulse tube design analysis. To acquire the needed data, the DRS flexure-bearing compressor was tested at Lockheed Martin ATC with dead volumes to determine its characteristics. Tests were performed using three different dead volumes (12. 1cc, 22.8cc, and 28.9cc), and five different charge pressures (300psia, 200psia, 100psia, 50psia, and 5psia). The compressor was driven with a current controlled amplifier with an input signal given by an HP signal generator. The power to the compressor was monitored with a Valhalla power meter, and a calibrated pressure transducer, mounted in the compression space, monitored the pressure. The resonant frequency at each charge pressure was determined by a frequency sweep searching for the maximum voltage for a fixed drive current, for drive currents ranging from 0.1 A up to 1.3 A. At the resonant frequency, the current, voltage, power, and pressure amplitudes were recorded. Table 1 presents a summary of the compressor parameters (for each compressor half); most were determined from the measurements, while some were provided by DRS. Compressor internal losses were also characterized to allow estimation of the expected efficiency of the overall pulse tube cryocooler. Because the compressor exit-passage parameters were designed for the standard DRS split-Stirling expander that has a small-diameter transfer line, somewhat higher losses were predicted when used with the pulse tube, which requires a larger transfer line. In the future, if more optimum performance from the compressor is desired, one should consider enlarging the internal flow passages to tailor the compressor for improved operation with a pulse tube.

Pulse Tube Design To achieve an efficient design for the pulse tube, detailed thermodynamic simulations were conducted by Lockheed Martin of the entire cooler system. Key parameters included pulse tube geometries, transfer line diameter, fill pressure, operating frequency, piston stroke, and pulse tube reservoir-line tuning.

The resulting design was predicted to provide 1.2 W of cooling at 80 K with 30 W of total compressor power and a piston amplitude of 2.9 mm. Note that the piston amplitude is well below the maximum of 5 mm. The predicted cooling capacity is slightly higher than the required 1.1 W, and the predicted specific power of 25 W/W matches the design goal. The largest uncertainty in the prediction was the internal losses within the compressor, which, in the dead volume tests, were particularly significant at high piston amplitudes. A conservative empirical model was used to represent the compressor flow losses in the analysis, which tended to reduce the piston amplitude in order to reduce the losses. A series of parametric studies was performed to predict the sensitivity of the coldhead to operating conditions. The efficiency of the coldhead was found to be relatively insensitive to mass flow rates and frequencies, typical of other L-M pulse tubes. This indicates that the coldhead design was not significantly influenced by the particular model used for the compressor losses. Figure 4 shows the (pressure-volume) PV specific power as a function of input power. As shown, the coldhead itself is predicted to have a PV specific power of 14 WAV at 80 K, comparable to other coldheads developed at Lockheed Martin ATC.5,6 Lockheed's best in-line, high-


Figure 4. Predicted PV specific power as a

159

Figure 5. Predicted cooling power dependence

function of input power.

on charge pressure.

capacity designs are typically 11-12 W/W. The slightly lower efficiency of this pulse tube is due in part to the U-tube configuration, the small diameter of the transferline internal to the compressor, and to the smaller cooling capacity. At the lower power levels, the efficiencies are still good,

although slightly decreased from the higher power levels. At 10 W of PV power, the PV specific power is predicted to be around 17.5 W/W. Figure 5 shows the predicted cooling power as a function of charge pressure at 30 W of compressor power. The frequency was varied for optimum performance, along with retiming of

the impedances. The regenerator and pulse tube remained fixed in the analysis. This plot suggests that 400 psia would be a good working charge pressure, and shows that the cooler can tolerate a slight reduction in charge pressure down to 300 psia or so without serious reduction in performance. However, if the charge pressure is decreased down to 200 psia, the performance begins to seriously degrade. If the regenerator and pulse tube were to be redesigned for lower

charge pressures, then the performance degradation would not be as severe as shown in Fig. 5. The results of the overall simulation analyses indicated that the proposed coldhead should perform well over a range of conditions. This is significant in that the compressor loss mechanisms were not known in detail. Once the compressor and coldhead were integrated, it was expected that minor tuning of the overall system would be able to achieve a good match between coldhead and compressor. Based on the modeling it was considered likely that the pulse tube would exceed the predicted

efficiency, since a conservative model was used for the compressor losses, a conservative model was used for the motor force constant, and many Lockheed coldheads outperform their predictions. Thus, it was expected that the coldhead would provide in excess of 1.1 W of cooling at better than 25 W/W. In addition, the low design stroke would allow the cooler to be driven to substantially higher strokes and power levels, although at a somewhat higher specific power.

Pulse Tube Fabrication Once the analyses and component designs were complete, the pulse tube cooler components were fabricated and assembled into a completed pulse tube expander. Figure 6 shows the piece parts ready for assembly, together with a completed pulse tube. Figure 7 shows the complete

cooler setup during verification testing at Lockheed. PULSE TUBE SYSTEM-LEVEL TESTING

After initial checkout and performance verification of the completed cooler at Lockheed Martin ATC, extensive performance characterization testing was carried out at JPL in preparation

for planned tests to validate the vibration and EMI compatibility with an actual gamma-ray detector using the setup illustrated in Fig. 3. Figure 8 presents the overall thermal performance measured at JPL as a function of coldend

160


Figure 6. Pulse tube expander piece parts and final assembly.

Figure 7. Completed pulse tube cooler with DRS compressor on the left, reservoir volume on the right, and pulse tube with vacuum bonnet assembly in the center.

Figure 8. Refrigeration performance of the completed gamma-ray pulse tube cooler as a function of input drive voltage, coldtip temperature, and coldtip load.


161

Figure 9. Refrigeration performance of the completed gamma-ray pulse tube cooler as a function of helium fill pressure.

temperature, coldend load, and input voltage (which is roughly proportional to stroke). Note that the specific-power performance at 80 K is around 22 W/W, which is better than the design goal of 25 W/W, and that the overall cooler capacity is also better than the design requirement, reaching over 1.6 watts at 80 K near full stroke (9 volts), in contrast to a requirement of 1.1 watt. To confirm the cooler's predicted sensitivity to fill pressure and drive frequency, additional parametric testing was conducted with these parameters as variables. The measured performance, displayed in Figs. 9 and 10, confirm that fill pressure increases cooling capacity with minimal effect on efficiency, while drive frequency, once the pulse tube volumes are fixed, is a relatively

sensitive parameter. For the as-fabricated pulse tube cooler, the best specific power is seen to occur at a frequency of around 42 Hz.

Figure 10. Refrigeration performance of the completed gamma-ray pulse tube cooler as a function of drive frequency.

162


SUMMARY AND CONCLUSIONS This paper has described the development, test, and performance of a novel new low-cost, low-noise, high-reliability pulse tube cooler, designed specifically for highly cost-constrained long-life space missions such as planetary gamma-ray spectroscopy. The developed cooler marries two technologies: a low-cost, high-reliability linear compressor and drive electronics from the 1.75 W tactical Stirling cryocooler of DRS Infrared Technologies, and an 80 K pulse tube developed specifically for the compressor by Lockheed Martin ATC. To achieve maximum life and low vibration, the compressor incorporates flat flexure springs for piston support and uses two opposing pistons in a head-to-head configuration with linear drive motors. The pulse tube is a compact U-tube configuration for improved integration and is mounted to the compressor in a split configuration with a transfer line. The successful new cooler achieves over 1.6 watts of cooling at 80 K at 23 W/W, and has the advantage of greatly reduced vibration at the coldtip and no life-limiting moving cold elements.

ACKNOWLEDGMENT

The work described in this paper was carried out by the Jet Propulsion Laboratory, California Institute of Technology, and by Lockheed Martin ATC under contract with JPL; it was sponsored via the Planetary Instrument Definition and Development Program (PIDDP) through an agreement with the National Aeronautics and Space Administration. REFERENCES 1. Rawlings, R.M. and Miskimins, S.M., “Flexure Springs Applied to Low Cost Linear Drive Cryocoolers,” Cryocoolers 11, Kluwer Academic/Plenum Publishers, NY, 2001. 2. Glaser, R.J., Ross, R.G., Jr. and Johnson, D.L., “STRV Cryocooler Tip Motion Suppression”, Cryocoolers 8, Plenum Publishing Corp., New York, 1995, pp. 455-463. 3. Ross, R.G., Jr., “JPL Cryocooler Development and Test Program: A 10-year Overview,” Proceedings of the 1999 IEEE Aerospace Conference, Snowmass, Colorado, Cat. No. 99TH8403C, ISBN 07803-5427-3, 1999, p. 5. 4. Johnson D.L., “Thermal Performance of the Texas Instruments 1-W Linear Drive Cryocooler,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, NY, 1999, pp. 95-104. 5. Kotsubo, V.,Olson, J.R., andNast.T.C., “Development of a 2W at 60K Pulse Tube Cryocooler for

Spaceborne Operation,” Cryocoolers 10, Kluwer Academic/Plenum Publishers, NY, 1999, pp. 157170. 6. Kotsubo, V., Olson,J.R., Champagne, P., Williams, B., Clappier, B. and Nast, T.C., “Development of Pulse Tube Cryocoolers for HTS Satellite Communications,” Cryocoolers 10, Kluwer Academic/ Plenum Publishing Corp., NY, 1999, pp. 171-179.

High Efficiency Pulse Tube Cooler E. Tward, C. K. Chan, J. Raab, T. Nguyen, R. Colbert and T. Davis†

TRW, Redondo Beach, CA 90278 † Air Force Research Laboratory Albuquerque, NM 87117

ABSTRACT

The High Efficiency Cooler (HEC) is being developed in order to provide a long life, low mass, high efficiency space cryocooler suitable for use on lightweight gimbaled optics on surveillance missions such as SBIRS Low. This paper reports on the development and testing of this next generation family of space pulse tube cryocoolers which feature high cooling capacity, lower mass,

lower EMI and lower self induced vibration than the current state of the art. The HEC achieves low input power and large cooling power because of the efficiency of its pulse tube cold head and highly efficient compressor. The low mass (10 watt

cooling capacity. The best previous specific mass (ratio of mechanical cooler mass to cooling power at 95 K) was in the range of 1.5 kg/W. This project seeks to make a major improvement to the specific power while at the same time reducing the cooler specific mass by 350% to 20,000 hours. PERFORMANCE CHARACTERISTICS OF PT405 Cool down performance

To maintain the standard specifications for both the 50 Hz and 60 Hz electrical power supplies, the PT405 is supplied with different compressor packages: the CP950 for 60 Hz, and the CP970 for 50 Hz. The input power in both cases is 4.9 kW with the 1st stage operating at 65 K and the 2nd stage at 4.2 K. A typical cooling-load map for the PT405 is given in Figure 2; for example, it shows that the

cooler provides 0.6 W at 4.2 K on the second stage, simultaneously with 30 W at 65 K on the first stage. The performance of the PT405 has been improved upward from the initial announced capacities of 0.57 W at 4.2 K and 18 W at 65 K.2

Figure 2. Typical cooling-load map of PT405.

Figure 1. Photo of PT405 cold head.

PERFORMANCE OF A 4 K PT IN CURRENT APPLICATIONS

207

Figure 3. Cool down curves for PT405 with and without mass attached.

The cool down time with a thermal mass attached for the PT405 was tested with 6.4 kg of OFHC copper on the 1st stage and 0.9 kg of OFHC copper on the 2nd stage. Figure 3 shows the

cool down times of the PT405 with and without the added mass. It takes 60 minutes for a unloaded PT405 to reach 4 K and 110 minutes for both stages to reach their minimum

temperatures of 2.6 K and 35 K. With the above thermal loads attached to both stages, the 2nd stage reaches 4 K in 100 minutes, and the two stages reach their minimum temperatures in 170 minutes.

Vibration of PT405

Cryomech, Inc. has collaborated with several different groups to analyze the vibrations generated by the PT405. Figure 4 shows the test rig used by GE R&D Center to analyze the mounting forces exerted on the cryostat at the PT405 base plate. A load cell was mounted under the base plate of cold head. The displacement of the 2nd stage cold heat exchanger was measured with both an optical comparator and an accelerometer. The accelerometer was mounted on the bottom of the 2nd stage cold heat exchanger as shown in Figure 4. The optical measurement was made with a ST405 Cryostat (Figure 5). A laser beam penetrated through the optical windows in the cryostat and focused on the second stage heat exchanger, measuring the movement. The mounting force of PT405 compared with that of a mechanical driven GM

cryorefrigerator is shown in Figure 6. The mounting force of PT405 is less than 3.6 N (0.8 lb) and for the GM 178 N (40 lb). The mounting force of pulse tube is only generated by the pressurization and depressurization of helium in the pulse tube assembly. The curve is similar in its wave shape to the dynamic gas pressures. Table 1 gives the displacements of the 2nd stage cold head measured by both the optical and the accelerometer methods. The vectors of the three-axis are given in Figure 4. The two tests measured similar displacements. The maximum displacement around is in the vertical direction.

208

GM-TYPE PULSE TUBE COOLERS FOR LOW TEMPERATURES

Figure 4. Vibration test rig .

Figure 5. Photo of ST405 cryostat.

PERFORMANCE OF A 4 K PT IN CURRENT APPLICATIONS

209

Figure 6. Mounting forces of pulse tube and GM cryorefrigerators.

Effects of a Magnetic Field or Orientation on the Cooling Performance

The effects of a magnetic field on the operation of the PT405 have been tested. Due to the limitation of resources, we have not tested the PT405 in a large enough magnetic field that has degraded its performance. We have designed the second stage regenerator with antiferromagnetic materials to minimize the loss of heat capacity in field. In one test we moved a PT405 Cold Head while operating in a cryostat at minimum temperature toward a superconducting magnet. In this test, a field strength of 900 gauss had no affect on the rotary valve motor or bottom temperature. Higher magnetic field will stop the motor. According to the published information of the rare earth materials, it is expected that the low temperature

performance of the 2nd stage will be slightly decreased in the magnetic field higher than 1 Tesla. It is suggested for maximum performance that the pulse tube cold head operate as close to vertical as possible. “Gas mixing losses” in the pulse tubes caused by gravity decrease the performance of both stages when the pulse tube cold head is tilted off the vertical position2. However, mounting angels up to 30° off vertical will not degrade the performance greatly.

210

GM-TYPE PULSE TUBE COOLERS FOR LOW TEMPERATURES

SOME APPLICATIONS OF PT405

Low vibration cryostat

A 3 K cryogen-free cryostat with optical access was built at Cryomech, Figure 5. The Model ST405 has very low vibration, as can be seen in Table 1. A radiation shield is fixed on the 1st cooling stage to reduce the radiation loss on the 2nd cooling stage and sample holder. It is designed for two and three axis optical measurements. There are 5 holes with diameter of 12.7 mm through the radiation shield and 5 optical windows of diameter of 25.4 mm on the optical cube. The radiation losses are crucial for the sample temperatures below 4 K. The bottom temperature on the sample holder in the cryostat is 2.6 K, when all holes through the radiation shield are covered with aluminum tape; 3.0 K with one hole uncovered; and 3.5 K with two holes uncovered. Small helium-4 liquefier or re-condenser

A small helium liquefier was developed at Cryomech using the PT405. In the system room temperature helium gas is pre-cooled by the 1st stage and 2nd stage regenerator. Figure 7 shows a schematic of the liquefier. The pre-cooling of the helium from the first stage temperature down to approximately 6.5 K using the inefficiency of the 2nd stage regenerator is critical for increasing the liquefaction rate. Numerical analysis made by the present authors calculates that the precooling heat load on the 2nd stage regenerator, decreases the PT405 2nd stage cooling capacity by only 10% of the heat actually absorbed into the regenerator. Figure 8 is a curve of the rate of liquefaction based on the liquid helium level in our condensation chamber. The accuracy of the level instrument was mm, and the liquid level sensor indicated that liquid helium level increased steadily. The pulse tube liquefier condensed 1.0 liters of liquid helium at 4.2 K in five hours, for a corresponding liquefaction rate of 4.8 liters per day.

Cryogen-free operation of superconducting magnet

The first commercial cryogen-free superconducting magnet system cooled by the PT405 was developed by Cryomagnetics, Inc. Figure 9 is a photo of the magnet system. The system has a horizontal room temperature bore diameter of 32 mm. The first stage of PT405 is used for cooling intermediate thermal shield and Bi-2223 based HTS current leads. The 2nd stage conductively cooled NbTi-based superconducting magnet below 4 K. The complete system takes approximately 14 hours to cool down the system to the operating temperatures: 50 K for

the thermal shield and 1275 K]). However, most of the commercially available “pure Er” metal already contain a substantial amount of interstitial impurities, and the interstitial alloy (noted below) was purchased from a commercial vendor without any additional alloying or processing. The presence of 2.7 at.% O, plus smaller amounts of N and C, 0.3 and 0.2 at.%, respectively, in pure Er, reduces the height of the sharp peak at 19 K and shifts it upward by about 3 K. Furthermore, it probably combines with the smaller 25 K peak in pure Er to give a small, but broader, heat capacity peak at 22 K (see Fig. 2). Furthermore, the 52 K peak is nearly eliminated while the 88 K peak is shifted downward to 84 K. As a whole the volumetric heat

MAGNETIC REGENERATOR ALLOYS FOR 10 TO 80 K RANGE

435

Figure 1. The volumetric heat capacity from 0 to 100 K of high purity Gd, Tb, Dy, Ho and Er metals,

along with those of bronze, stainless steel and lead. The letters SSE stand for solid state electrolysis, which was used to purify these metals.

capacity of pure Er between the various peaks is not changed significantly by the addition of the interstitial elements and is still better than lead over 20 to 80 K region. Throughout the rest of the paper this alloy will be used as the prototype Er alloy against which the heat capacities of the other Er-based alloys will be compared.

Figure 2. The volumetric heat capacity of an Er-based interstitial alloy

100 K. Also shown are the heat capacities of high purity Er and lead.

from 3.5 to

436

REGENERATOR ANALYSIS AND MATERIALS DEVELOPMENTS

Non-lanthanide Substitutional Alloys A five atomic percent addition of Sc or Y or Zr or Hf wiped-out the 19, 25 and 52 K peaks and lowered the 88 K one. Below 18 K the volumetric heat capacity is slightly larger by a few percent over that of the prototype alloys and Pb. But as a whole these alloys are not much of an improvement, if at all, over the Er-based interstitial alloy as a regenerator alloy. Lanthanide Substitutional Alloys

The heavy lanthanide elements (Gd, Tb, Dy and Ho) tend to raise the magnetic transitions temperature of Er, and thus these metals are not useful in increasing the heat capacity below 20 K. Lutetium, the last member of the lanthanide elements, has no unpaired 4f electrons, and its

behavior is similar to that observed for the non-lanthanide alloying agents Sc, Y, Zr and Hf, see above.

The light lanthanides (La, Ce, Pr and Nd), however, have an unusual influence on the magnetic ordering phenomena in Er. With respect to the upper (or Néel temperature) transition of Er, all four of these elements lower it, in proportion to their size: the larger La atom lowers 88 K peak the most (to ~70 K for 5 at.%), followed by Ce (about the same as La), which is

followed by Pr (to ~75K for 5 at.%) and Nd (to ~78 K at 5 at.%). For the lower temperature transformations, La and Ce behave quite similarly, and significantly different from that of Pr and

Nd, which have similar affects. For La and Ce small additions (5 at.%) wipe-out the 52 K transition while rapidly raising

and merging the 19 and 25 K peaks into one. The 19/25 K peak eventually merges with the rapidly dropping Néel temperature (88 K) at 20 at.% (La or Ce). This results in large somewhat broad (11 K) heat capacity peak at ~40 K, but the heat capacity below 34 K and above 45 K lies below that of the alloy. The maximum heat capacity is 1.8 J/cm3K at 40 K for the and and the alloys would be useful if one needed a high heat capacity material for the 35 to 45 K range. The addition of 5 at.% Pr or Nd destroys the 25 K heat capacity anomaly shifting the entropy toward the 19 K Curie temperature peak while reusing its temperature to about 23 K.

The 52 K peak is rapidly lowered by the initial additions of Pr reaching a minimum at ~7 at.% Pr before rising slightly and leveling-off when more than 10 at.% Pr is added. As more

Pr (or Nd) are added to the Er (~15 at.%), a double peak structure develops, which eventually merges into one peak at ~27 at.% Pr, due to rapidly dropping of the 88 K transition temperature. These behaviors are shown in Fig. 3a which shows the change of the transition temperatures and Fig. 3b which shows the associated heat capacity peak values, both as a function of Pr content. The heat capacities of a series of Er-Pr alloys are shown in Fig. 4. It is noted that for alloys

containing more than 30 at.% Pr the heat capacity is larger than that of either Pb below 20 K. This is shown in more detail in Fig. 5, where it is seen that highest heat capacity of any of the Er-Pr alloys and the Er interstitial alloy

Pb or Er3Ni between 10 and 20 K. At 10 K the

or has the

alloy’s heat capacity is 185% larger than

that of Pb and the same as that of Er3Ni.

The Nd additions follow a similar trend as shown by the Pr addition, but since the upper transition is lowered at a slower rate, more than 30 at.% Nd is required before the two peaks merge into one. Based on these results we suggest the following combination of Er-based alloys as a replacement of lead in the cryocooler regenerator: at the high temperature end the interstitial

alloy; the

alloy for the intermediate temperature region; and the

alloy for the low temperature end of the regenerator. The heat capacities of these three

Er-based alloys are shown in Fig. 6 along with that of lead. It is evident from Fig. 6 that the alloy would be the best regenerator material from 40 to 85 K having a heat capacity 20 to 40% larger than that of Pb; while which has a heat capacity 20 to 30% larger than that of Pb, would cover the 24 to 40 K range, and would be the most efficient below 24 K, since its heat capacity is 20 to 185% larger than lead. Of course if cooling down


437

Figure 3. The variation of the magnetic ordering temperature (a) and the peak value of the heat capacity at the respective ordering temperature (b) as a function of the Pr concentration.

much below 10 K (i.e. down to 4 K) is required, a lower temperature stage regenerator composed of Er3Ni or HoCu2 or Nd would be needed, as is the standard practice today, when Pb is used as the intermediate temperature stage regenerator material.

Figure 4. The volumetric heat capacity of a series of Er-Pr alloys from 3.5 to 100 K and the prototype interstitial alloy and Pb.

438


Figure 5. The volumetric heat capacity of some Er-based alloys from 3 to 22 K. The heat capacities of Er3Ni and Pb are shown for comparison.

Other Properties In addition to the significantly higher heat capacity, these three Er alloys have other distinct advantages over lead with a lower thermal conductivity and a higher strength; and they are not as toxic as Pb, which is a federally regulated poison. The thermal conductivity which is shown in Fig. 7, is nearly one order of magnitude lower than that of Pb, comparable to that of stainless steel. The reduced thermal conductivity would lead to lower longitudinal heat losses in the

Figure 6. The heat capacities of

and

along with that of Pb.


439

Figure 7. The thermal conductivities for interstitial Er, Pb and stainless steel.

regenerator. The thermal conductivity of the slightly lower than that of the Pr atoms in the Er matrix.

Er73Pr27 and Er50Pr50 would be expected to be because of additional phonon scattering from the

The ultimate tensile strength of interstitital Er, Pb, stainless steel and bronze are listed in

Table 1. It is evident that the tensile strength of is an order of magnitude larger than that of Pb, and about the same as bronze. This is important if the Er-based alloy are used in the form of spheres, since the materials are strong enough to prevent the loss of sphericity as can happen with Pb (Sb-hardened) alloys. The strength of the Pr-Er alloys would be expected to be about the same as that of the interstitial Er alloy. This level of strength is important when these Er-based alloys are used as wires, screens, flat sheets and jellyrolls in cryocooler regenerators, which one cannot do with Pb since it is so weak.

There are two other important features which need to be mentioned. One is that the Er alloys do not oxidize like the light lanthanides, e.g. Nd, which has been used in cryocooler regenerators. We have held and Er60Pr40 at 396±5 K for over 15 months and there has been no measurable weight gain or loss (within ±0.1 mg) for samples weighing 3.6, 2.5 and 4.1 g, respectively. This is in contrast to Nd metal which oxidizes to form Nd2O3 within a few hours after being exposed to ambient air at room temperature. Thus these alloys can be easily handled and stored without any special precautions. The second point is that since the Er-based alloys are solid solution alloys, and not intermetallic compounds, they can be readily fabricated into spheres, wires and foils (ribbons), see Fig. 8. Since they are ductile alloys with reasonable strength they will not decrepitate or crumble, which can easily occur when using brittle intermetallic compounds in regenerators due

440


Figure 8. Fabricated forms of spherical powders – 0.30mm (12 mil) diameter; ribbon – 0.05mm (2 mil) thick; and wire – 0.30mm (12 mil) diameter.

to the high pressure pulses of the gases that are rapidly recycled through the regenerator. Details on the fabrication of the and alloys into spherical powders are being presented in another paper at this conference.12

Initial Test Results An initial test using in place of Pb as the regenerator material in a single stage pulse tube cryocooler indicated that at low frequencies the Er interstitial alloy performed better than Pb. Additional results using conference.13

are presented in another paper at this

CONCLUSIONS AND SUMMARY

We have shown that Er-based alloys in comparison to Pb have significantly (1) higher volumetric heat capacities from 10 to 80 K, (2) lower thermal conductivities, and (3) improved tensile strength. The Er alloys are oxidation resistant below 396 K (123°C), and can readily be formed into wires, ribbons (foils) and spheres. We have suggested that the most efficient replacement for lead in a regenerator is to use a combination of three alloys at the hot end, at the intermediate temperatures, and at the cold end. Finally we wish to note that these Er-based alloys are not a replacement for Er3Ni, Nd or HoCu2, which are required to reach temperatures below 10 K, but are to be used in conjunction with them if temperatures below 8 K are required. Our theme is not only to improve cryocooler performance

but also to “get the lead out.” ACKNOWLEDGEMENTS This work was supported in part by Atlas Scientific, Sunnyvale, California via a SBIR, and in part by the Materials Sciences Division, Office of Basic Energy Sciences, U.S. Department of

Energy under contract W-7405-ENG-82.

MAGNETIC REGENERATOR ALLOYS FOR10 TO 80 K RANGE

441

REFERENCES

1. Buschow, K.H.J., Olijhoek, J.F. and Miedema, A.R., “Extremely Large Heat Capacities Between 4 and 10 K”, Cryogenics, vol. 15 (1975), pp. 261-264. 2. Sahashi, M., Tokai, Y., Kuriyama, T., Nakagome, H., Li, R., Ogawa, M. and Hashimoto, T., “New Magnetic Material R3T System with Extremely Large Heat Capacities Used as Heat Regenerators”, Adv. Cryogen. Eng., vol. 35 (1990), pp. 1175-1182.

3. Kuriyama, T., Hakamada, R., Nakagome, H., Tokai, Y., Sahashi, M., and Li, R., Yoshida, O., Matsumoto, K. and Hashimoto T., “High Efficient Two-Stage GM Refrigerator with Magnetic Materials in Liquid Helium Temperature Region”, Adv. Cryogen. Eng., vol. 35 (1990), pp. 12611269. 4. Gschneidner, K.A., Jr., “Physical Properties and Interrelationships of Metallic and Semimetallic Elements”, Solid State Phys., vol. 16 (1964), pp. 275-426. 5. Although the crystalline electric field splitting of the Er 4f electrons in the ground state multiple 4I15/2

in Er3Ni are not known, those of the isostructural Er3Co compound have been determined. The calculated and measured Schottky heat capacity of Er3Co shows a broad peak at ~12 K which slowly falls off with increasing temperature reaching about half the peak value at 30 K. We would expect that the Er3Ni heat capacity would show a very similar behavior. This work was reported by: Takahashi Saito, A., Tutai, A., Sahashi and M., Hasimoto, T., “Crystal Field Effects on Thermal and Magnetic Properties of Er3Co”, Jpn. J. Appl. Phys., vol. 34 (1995), pp. L171-L173. 6. Merida, W.R. and Barclay, J.A., “Monolithic Regenerator Technology for Low Temperature (4 K) Gifford-McMahon Cryocoolers”, Adv. Cryogen. Eng., vol. 43 (1998), pp. 1597-1604. 7. Sinha, S.K., “Magnetic Structures and Inelastic Neutron Scattering: Metals, Alloys and Compounds” in Handbook on the Physics and Chemistry of Rare Earth, Gschneidner, K.A., Jr. and Eyring, L., eds., North-Holland Publishing Co., Amsterdam (1978), vol. 1, chap. 7, pp. 489-589. 8. McEwen, K.A., “Magnetic and Transport Properties of the Rare Earths” in Handbook on the Physics

and Chemistry of Rare Earth, Gschneidner, K.A., Jr. and Eyring, L., eds., North-Holland Publishing Co., Amsterdam (1978), vol. 1, chap. 6, pp. 411-488. 9. Scott, T.E., “Elastic and Mechanical Properties” in Handbook on Physics and Chemistry of Rare

Earths, Gschneidner, K.A, Jr. and Eyring, L., eds., North-Holland Publishing Co., Amsterdam (1978), vol. 1, chap. 8, pp. 591-705. 10. Anonymous in Properties and Selection: Nonferrous Alloys and Special-Purpose Materials, Metals Handbook, 10th ed., ASM Intern., Materials Park, OH (1990), vol. 2, p. 217 (Bronze) and p. 550 (Pb). 11. Anonymous in Properties and Selection: Irons, Steels and High-Performance Alloys, Metals

Handbook, 10th ed., ASM Intern., Materials Park, OH (1990), vol. 1, p. 855.

12. Miller, S.A., Nicholson, J. D., Gschneidner, K.A., Jr., Pecharsky, A.O. and Pecharsky, V.K., “Manufacturing Considerations for Rare Earth Powders Used in Cryocooler and Magnetic

Refrigerator Applications”, paper presented at 11th International Cryocooler Conference, June 20-22, 2000, Keystone, Colorado.

13. Kashani, A., Helvensteijn, B.P.M., Gschneidner, K.A., Jr., Pecharsky, A.O. and Pecharsky, V.K., “New Regenerator Materials for Use in Pulse Tube Coolers”, paper presented at 11th International Cryocooler Conference, June 20-22, 2000, Keystone, Colorado.

Low Temperature Properties of HoSb, DySb, and GdSb H. Nakane, S. Yamazaki, H. Fujishiro*, T. Yamaguchi**,

S. Yoshizawa**, T. Numazawa*** and M. Okamura****

Kogakuin University, Shinjuku-ku, Tokyo, 163-8677, Japan * Iwate University, Ueda, Morioka, 020-8551, Japan ** Meisei University, Hino-shi, Tokyo, 191-8506, Japan ***

National Research Institute for Metal, Tsukuba-shi, 305-0003, Japan

****

Toshiba Corporation, Yokohama-shi, 235-8522, Japan

ABSTRACT

Materials including rare-earth and Sb compounds were developed as regenerator materials for a 4K GM refrigerator. Sb compounds have a remarkably large peak value of specific heat in the low temperature range. First, the heat capacities, thermal conductivity, and thermal expansion of HoSb, DySb and GdSb, which are anti-ferromagnetic with a large spin value in the low temperature range, were measured. Then, the results of these measurements were used to analyze the problems encountered when Sb compounds were used as the regenerative materials in a GM refrigerator. As for the thermal expansion, the measured values of Sb compounds were compared with that of fabric-impregnated phenol-formaldehyde resin, which is usually used as the wall material of a regenerator. As regards the thermal conductivity, the thermal diffusivity was obtained by using the

measured values of thermal conductivity and specific heat. The heat penetration depth was evaluated from the thermal diffusivity. Discussion of the heat penetration depth is based on the figures obtained in a previous experiment using a GM refrigerator with HoSb packed at the cold end of the second stage of a multi-layer regenerator. INTRODUCTION

In a small 4 K refrigerator, the refrigeration capacity depends on the response speed of heat exchange, i.e., the exhaustion and absorption of heat between the working gas and the materials packed into the regenerator. In order to obtain higher effectiveness in the regenerator, the heat capacity of the regenerator materials must be larger than that of 4He used as the working gas. Only a magnetic material that has a large specific heat based on magnetic phase transition is effective below 15 K. However, the specific heat of pressurized 4He gas is very large below 15 K. Since the heat exchange region of 4He gas is wide, a single magnetic material cannot cover the specific heat of4He gas. Multi-layer regenerators composed of several rare-earth compounds that have different specific heat peak temperatures are usually used. HoSb and DySb compounds were discovered to have a remarkably large specific heat peak. HoSb compound has the peak value of 2.7 J/K cm3 at 5.3 K, and DySb, 2.0 J/Kcm 3 at 9.5 K. The measured results of the specific heat1 of HoSb, DySb and GdSb are shown in Fig. 1. HoSb was Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Temperature dependence of specific heat of Sb compound.

packed into a multi-layer regenerator and the refrigeration capacities were measured and are included in the report.2

In regards to the measured values of thermal expansion, the relation between the fabric-impregnated phenol-formaldehyde resin (fabric-impregnated Bakelite), which is usually used as the wall material of regenerators, and Sb materials, which have a remarkably large peak, was analyzed in this experiment. Then, the thermal conductivities of HoSb, DySb and GdSb were measured, and the thermal diffusivity was obtained from the measured values of thermal conductivity and the specific heat. The heat penetration depth of HoSb was compared with that of4He used as the working gas. Whether the driving cycle of the displacer was optimal was verified for the former experiment2 in which HoSb was packed at the cold end of the second stage multi-layer regenerator of a GM refrigerator.

EXPERIMENTS The samples used were intermetallic compounds of HoSb, DySb and GdSb made by arc welding at the melting points of 2433 K (2160°C), 2443 K (2170°C) and 2403 K (2130°C), respectively. Two Sb compounds were found to have remarkably large specific heat peak values in the low temperature range. One of these compounds is HoSb, with the peak value of 2.7 J/K cm3 at 5.3 K, and the other one is DySb, with 2.0 J/K cm3 at 9.5 K. GdSb has a broad specific heat peak around 24 K. The measuring methods of thermal conductivity and thermal expansion of such magnetic materials are explained and the measuring results are shown as follows:

Measurement and Discussion of Thermal Expansion The thermal expansions of HoSb, DySb and GdSb were measured by a clip type dilatometer as shown in Fig. 2. A resistance bridge was formed by four strain gauge elements (Kyowa Electronic Instruments, KFL-1-120-C1-11) bound to a phosphor-bronze clip. The balance of the resistance bridge was directly measured to an accuracy of 71/2 figures with an automatic thermometer bridge (TINSLEY 5840D) by the four-wire method. At a certain temperature, the balance of the resistance bridge changes almost linearly to the length of the specimen. To obtain the calibration values of for the expansion, the of two copper reference specimens of different lengths was measured, as the thermal expansion rate of copper is well known.

Figure 2. Schematic diagram of the clip type dilatometer using four strain gage elements.

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Figure 3. Thermal expansions of GdSb, DySb and HoSb.

Figure 4. Temperature dependence of Sb compounds and Bakelite.

The lengths of HoSb, DySb and GdSb samples were 10 mm, and the cross sections about 6.3 mm x 2.3 mm. The accuracy of the measurement increases with the size of the cross section. As for the fabric-impregnated phenol-formaldehyde resin (fabric-impregnated Bakelite), which is usually

used as the wall material of regenerators, a sample of 10 mm in diameter and 10 mm in length was cut out from a circular rod. The linear expansions of HoSb, DySb and GdSb are shown in Fig. 3. As shown in Fig. 1, HoSb had the highest specific heat peak, DySb second, and GdSb third. However, the sharpness of change in the linear expansion around the Néel temperature was DySb, first; HoSb,

second; and GdSb, third (not sharp). In Fig. 4, the linear expansions normalized at 300 K are expressed as percentages. The linear expansions of the samples at 4 K were about – 0.2 % for Sb compounds, – 0.5 % for Bakelite against the diameter and – 0.2 % against the length. The contraction of Bakelite in diameter was larger than those of HoSb and DySb. In a former experiment,2 in which HoSb was packed at the cold end of the second stage of a multi-layer regenerator of a GM refrigerator, the effect of contraction of the wall materials was compensated by using felt as the partition material.

Measurement and Discussion of Thermal Conductivity The thermal conductivity was measured by the steady state heat flow method. Fig. 5 shows a schematic view of the sample set on top of the cold head of a GM refrigerator used as a cryostat3. Both samples were glued to the cold-finger of the refrigerator and the metal film resistance heater (10 kW) to the sample with GE7031 varnish. AuFe(0.07 at.%)-Chromel thermocouples with a di-

ameter of 73 µm were used as thermometers. The temperature range for the measurement was from 4 to 300 K. Air from the whole chamber of the sample was removed to 1 x 10–6 Torr with an oil diffusion pump. The samples were cut to the dimensions of 3 mm x 3 mm x 10 mm as the accuracy

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Figure 5. Schematic diagram of the thermal diffusivity measurement system.

Figure 6. Temperature dependence of thermal conductivity.

of measurement inversely increases with the size of the cross section in this method. The experimental results of the thermal conductivity of HoSb, DySb and GdSb are shown in Fig. 6. The thermal conductivity of lead (Pb), stainless steel, ErNi2, ErCo2 and DyNi2 which are used as conventional regenerator materials, is shown in Fig. 6 for comparison. Heat is periodically absorbed and exhausted by the materials used in a regenerator. Thermal diffusivity is often a more convenient parameter than thermal conductivity in discussing such a non-steady heat transfer subject. As is well known, thermal diffusivity D is given by: where is the thermal conductivity and Cp the volumetric specific heat. By substituting the temperature properties of the specific heat and thermal conductivity of each material into Eq. (1), the thermal diffusivities D are obtained as shown in Fig. 7. As shown in Fig. 7, the thermal diffusivities

of HoSb and DySb at the Néel temperature are very low in comparison with that of Pb. It is, therefore, necessary to discuss whether the materials are effective under the conditions of non-steady heat transfer that occurs in regenerators. Obviously the amplitude of temperature change is too small at

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Figure 7. Temperature dependence of thermal diffusivities of Sb compounds.

a position far from the surface to cause effective heat transfer between the surrounding gas and the material. Let us consider a one-dimensional semi-infinite body model, as discussed in literature.4 In this simple model, heat is transferred in only one direction and the body is heated periodically only on the surface. This model is often used to check the general tendency of the measured data.

We define the position where the amplitude of temperature change is 1 / e times that on the surface as the penetration depth of heat. The penetration depth of heat Ld is described as:

We assume that the material between the surface and Ld absorbs and exhausts the heat effectively, and suggest that the material dimensions used should be smaller than Ld. The heat penetration depth of three different cycles is shown as a function of thermal diffusivity in Fig. 8. The effective range of thermal diffusivity for HoSb, DySb, Er3Ni and lead is also shown. The optimum sizes of the regenerator materials can be estimated from Fig 8. Based on a previous experiment with a GM refrigerator2 in which the displacer was operated at a frequency of 74 rpm and the size of the regenerator material was 0.3 mm, we can estimate from Fig. 8 that either Sb or lead will be

effective in the heat exchange process. Figures 9 and 10 describe the temperature dependence of the heat penetration depth of HoSb with 4He as the working gas, as obtained from Eq. (1), and using the operating frequency as a parameter. When the displacer was operated at a frequency of 74 rpm, a shallow heat penetration depth was observed around the Néel temperature (5.3 K) and the penetration depth was 0.7 mm.

Figure 8. Relation between penetration depth of heat and thermal diffusivity.

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Figure 9. Thermal diffusion of HoSb.

Figure 10. Thermal diffusion depth of 4He at 1.5 MPa.

The effectiveness of heat exchange should not be affected by the size of the particles as 0.3 mm particles, smaller than 0.7 mm, were used in this experiment. The penetration depth of the heat of 4He over 0.3 mm at 600 rpm can be usable because it is considered larger than the hydraulic diameter.

CONCLUSION The basic properties of HoSb, DySb and GdSb as regenerator materials were measured at a low temperature. As for the thermal expansions, the relationship between the phenol-formaldehyde resin (Bakelite) and the Sb materials, which have a remarkably large peak, was analyzed. The thermal conductivity of HoSb, DySb and GdSb was also measured and the thermal diffusivity was furthermore obtained by using the measured values of the thermal conductivity and specific heat. The heat penetration depth was evaluated from the values of thermal diffusivity. It was compared with the thermal expansion of 4He gas. Verification of the optimal driving cycle for the displacer was carried out for a previous experiment in which HoSb was packed at the cold end of the second stage multi-layer regenerator of a GM refrigerator.

REFERENCES 1. Nakane, H. et al., “Multilayer Magnetic Regenerators with an Optimum Structure around 4.2K,” Cryocoolers 10, Plenum Press, New York (1999), pp. 611-620. 2. Nakane, H. et al., “Refrigeration Capacity of a GM Refrigerator Utilizing HoSb as its Regenerative Material,” Advances in Cryogenic Engineering, vol. 45, (2000), pp. 397-402. 3. Ikebe, M. et al., “Simultaneous Measurement of Thermal Diffusivity and Conductivity Applied to Bi-2223 Ceramic Superconductors”, Journal of the Physical Society of Japan, vol. 63, No. 8, (1994), pp. 3107-3114. 4. Ogawa, M. et al., “Thermal Conductivities of Magnetic Intermetallic Compounds for Cryogenic Regenerator,” Cryogenics, vol. 31 (1999), pp. 405-410.

Manufacturing Considerations for Rare Earth Powders Used in Cryocooler and Magnetic Refrigerator Applications S. A. Miller1, J. D. Nicholson1, K. A. Gschneidner, Jr.2,3, A. O. Pecharsky2, and V. K. Pecharsky2,3 1 Starmet Corporation Concord, MA 01742, USA 2 Ames Laboratory and 3Department of Materials Science and Engineering Iowa State University Ames, IA 50011-3020, USA

ABSTRACT The high chemical reactivity and oxygen affinity of the rare earth metals make them especially difficult to produce and maintain as high purity powders. From the initial oxide reduction through bulk melt preparation and finally powder production the sensitivity of the metals and alloys to contamination must be attended to at all times. Production of these powders has raised new challenges for the powder manufacturer. Process adaptation to the unique requirements of the rare earth materials to insure a quality product will be described. The physics of the current powder production technique will be reviewed and related to the actual product made. Individual processing steps with their influence on final product quality will be reviewed. Finally, the knowledge gained during the pre-production learning curve will be used to define potential manufacturing process that could be used for the economic production of large quantities (>100 kg) of these powders when needed. The volumetric heat capacities of powders of Gd and Er metals, and of an Er-Pr alloy, which were prepared by the Plasma Rotating Electrode Process (PREP) have been measured and compared to the bulk starting materials. The heat capacities are essentially the same within experimental error. The three materials have been used as regenerator materials: Gd for an active magnetic refrigerator, and Er and Er-Pr as passive regenerator materials for Gifford-McMahon and pulse tube cryocoolers, INTRODUCTION The rare earth powders used in this study were made by PREP®. Powder is produced by controllably melting the end of a round bar of the chosen material with a plasma torch while the bar is rapidly rotating about its longitudinal axis. The molten metal is centrifugally ejected from the surface of the bar and forms droplets that solidify to form spherical powder particles. A conceptual schematic of the process is shown in Figure 1. Cryocoolers 11, edited by R.G. Ross, Jr.

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Of the current commercially available metal atomization processes PREP has several inherent process characteristics that make it uniquely suitable for the fabrication of powders for

high performance applications. First, because PREP is a means of contactless melting and atomization, PREP produces powder with the highest level of cleanliness possible. This is a critical feature for reactive, high melting temperature alloys that are aggressively corrosive in their molten state and attack conventional ceramic crucibles and for alloys that have a high sensitivity to very small levels of contamination. Such alloys are routinely atomized by PREP without incurring contamination. Examples are titanium, zirconium, molybdenum and rare earth metals and alloys. Second, PREP powder is almost perfectly spherical and essentially satellite free, because during the atomization process droplets are dispersed and move radially away from each other. Thus, there is little opportunity for collisions between droplets and particles and the resulting coalescence of the two into irregularly shaped clusters. This single particle nature of the powder spheres results in PREP powder being very free flowing and having a high packing density, approximately 65%. Third, PREP powder combines both a tighter geometric standard deviation and a larger median size than can be produced via other techniques.1 Finally, because PREP atomization mechanism is produced by centrifugal forces rather than by aerodynamic drag the powder is essentially porosity free when compared to gas atomized powder. PREP is conducted in a vacuum/controlled atmosphere tank 2440 mm (96 in.) in diameter by 600 mm (24 in.) long. Tank dimensions limit maximum powder size to ~1.5 mm, while maximum spindle speeds of 2620 radians/s limit minimum powder size to ~40 µm. Usually, powder production is conducted under inert gas; the preferred medium is helium, which offers

both improved heat transfer properties and plasma performance.

Accurately controlling the rotation speed of the bar to be melted is necessary to obtain a desired median particle size. The diameter of the molten droplet diameter is determined by the properties of the liquid metal, the centrifugal ejecting forces (related to rotation speed), and to a limited extent the aerodynamics of the droplet trajectory through the inert cover gas. The following equation predicting median droplet size is obtained from a force balance of the centrifugal forces acting on the molten surface tending to cause atomization with the surface tension forces resisting atomization:

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where d is the median droplet diameter (microns); is rotation rate (radians/s); and are the surface tension (dynes/cm) and density (g/cm3) of the alloy being atomized, D is the electrode diameter (cm) and K is an empirical constant principally determined by the method of droplet formation which is in turn controlled by the melting rate.2 For any particular alloy the material properties are fixed and (1) can be further simplified to

where K is a constant (over a limited melting rate range) that has been determined for many alloy

systems.3 Lacking from current process understanding, however, is the ability to control the geometric standard deviation of the powder produced. PREP like the other commercial atomization processes is an inherently stochastic process, and the degree of spread in the particle

size distribution is outside operator control. While the amount of spread can not be controlled directly, it is noted that the geometric standard deviation for PREP is approximately 1.2, providing a significantly tighter particle size distribution than competitive processes. RESULTS AND DISCUSSION

The Pr, Gd and Er metals were purchased as chunks from Tianjiao International Trading Company (formerly Baotou Steel and Rare Earth Company) of China. The vendor stated that the

metals were 99.9% pure. Chemical analysis indicated that the Pr metal was only 97.8a/o (99.5w/o) pure (the major impurities were [in a/o] 0.60 O, 0.4 Li, 0.3 Al, 0.3 Fe, 0.2 Si, 0.2 C, 0.2 Nd; Gd 93a/o pure (99.7w/o) pure (the major impurities were [in a/o] 3.2 H, 1.2 O, 1.0 C, 0.8 F, 0.4 N and 0.1 Ca); and Er 96.8a/o (99.4w/o) pure (the major impurities were [in a/o] 2.7 O,

0.7 F, 0.4 H, 0.3 N, 0.2 C, 0.2 Ta, 0.1 Ca). Five to 15 kg of the metal chunks were vacuummelted at the Ames Laboratory in a tantalum crucible at 1600°C for Pr and Gd (1700°C for Er) for one hour to reduce calcium and hydrogen impurities by vaporization. The molten metal was bottom-poured to form 6.5 cm diameter rods. The rods were finish machined in preparation for PREP processing. In the case of Gd for the first run, 26% of the initial 13.6 kg of metal yielded spheres with a diameter < 150 µm, 32% of the spheres between 150 and 300 µm, and 5% of the spheres with a diameter > 300 µm. The other major losses of material (other than the unwanted size spheres were volatiles in the melting operation (minor), slag in the casting step (intermediate), and machining (major) and the end stubs from PREP process (intermediate). The yield of usable spheres (between 150 and 300 µm) from a second run of Gd metal was increased to 51%. Chemical analysis of the spheres revealed that the hydrogen content was reduced to 0.08 a/o and the calcium to < 0.005 a/o, while the other interstitial impurity contents remained essentially the same. Similar results were obtained on the Er materials. The Pr-73a/oEr was prepared by melting chunks of weighed amounts of Pr and Er in the proper ratio in a Ta crucible under an Ar atmosphere at ~1565°C before drop casting the molten alloy into a water cooled Cu crucible. The melting point of this alloy was estimated to be 1440°C. The Pr to Er ratio was determined by x-ray fluorescence analysis along the length of the rod and a cross-section perpendicular to the rod axis. Since the Pr concentration (and thus the Er content) varied by more than ±1 a/o, the sample was broken up and remelted. After remelting the concentration of Pr was uniform to with ±1 a/o. After machining, this alloy was used to make powders by the PREP process. It is important that the Pr concentration be uniform

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throughout the rod, because the low temperature heat capacity varies significantly with changing

Pr concentration.4 If the Pr concentration is not reasonably uniform the produced powders will have a range of compositions, and thus, a range of low temperature heat capacity values.

Furthermore, these powders of variable composition will be intimately mixed and impossible to separate into various fractions with more uniform compositions and properties. The results for the Gd powders used as an active magnetic regenerator in a magnetic refrigerator were previously reported in 1997.6 The 1997 results showed that magnetic refrigeration is a viable cooling technology and competitive with conventional gas compression systems. In addition it is an environmentally friendly technology eliminating CFC, HCFC and NH3 gases. Before each powder type was atomized the vessel was evacuated to less than 50 µm and then backfilled with helium to a positive pressure of 40 kPa. The bars were brought up to the desired rotational speed and melting was commenced. Once atomization was completed the heat source was shut off, the powder allowed to cool to ambient and then the powder was removed from the chamber under an inert cover. A sample for sieve analysis was removed, and the remainder of the powder sized to specification for testing. The powder was then packed and shipped under argon. Further analysis was performed on the Er and Pr-a/o73 powders at the Ames Laboratory and is reported elsewhere in this proceedings.4 Figure 2 is a SEM micrograph of the Pr-73a/oEr powder produced in this study. Already classified to –250+104 µm the micrograph demonstrates the high level of sphericity and the lack of satelliting that characterizes PREP powder. While the physics governing the PREP process is

well understood and easily mathematically reduced to equation 1 there is a lack of sufficient melt property data, to make full use of equation 1 in a production environment. A fair amount exists for elemental materials, but as Table 1 implies, the data is scarce for binary and more complex alloys. As a result frequently equation 2 is employed at the expense of having to make a prototype run in order to determine the constant K before making large quantities of powder.

Thus the rotational speed, in Table 2 was initially back calculated by using equation 1 with the limited data and a known desired particle size d. Due to the scarcity of the required data, the actual speed shown in Table 2 is a compromise between operator experience and the predicted

Figure 2. SEM micrograph of Pr-73a/oEr powder showing the high degree of sphericity and lack of satelliting.


Figure 3. Erbium powder size distributions produced by the PREP process.

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value. The powder from this run is then analyzed for the median size produced, that value is entered into equation two with the speed and bar diameter employed and the process constant K is back calculated. Equation 2 is then used with the desired median particle size to determine the correct rotational speed for an extended production run. Figure 3 shows the particle size distributions obtained by this procedure for Er and how rapidly the process can converge on a desired size. The low temperature volumetric heat capacity of the Pr-73a/oEr alloy before and after processing the alloy into spheres is shown in Figure 4. As is evident the heat capacity of the Pr73a/oEr spheres are essentially the same as that of the bulk original starting alloy. The slight difference in the heat capacities at the 36 K peak is due to the fact that the spherical

Pr-73a/oEr powders are mixed with Ag powders (~30 volume %) and compacted in order to make the heat capacity measurements. When the known heat capacity of Ag is subtracted from the measured value this can result in a slight subtraction error especially at temperatures above

~30 K where the heat capacity of Ag increases significantly relative to that at lower temperatures. CONCLUSION

Given the current level of demand for monosized, or near monosized rare earth powders for cryocooler and magnetic refrigerator applications PREP processing offers the best current

commercial product in terms of sphericity, lack of contamination and tight size distribution. With careful manufacturing practice powder can be produced without a loss in the properties

relevant to cryocoolers and magnetic refrigerators. However, as demand for the powder increases the development of newer technologies to production status will be warranted to provide improved product to the cryocooler community. From the powder producer's point of

Figure 4. The volumetric heat capacity from 4 to 79 K for both the bulk form and 100-250 µm spheres of Pr-73a/oEr obtained by the PREP process.


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view the remaining challenge is to produce a more monosized spherical powder for improved material utilization and system performance. ACKNOWLEDGEMENTS

Ames Laboratory is operated for the U.S. Department of Energy (DOE) by Iowa State University under Contract No. W-7405-ENG-82. The work was partially supported by the Advanced Energy Projects and Technology Research Division, Office of Computational and Technology Research of DOE; by Atlas Scientific, Sunnyvale, California via a SBIR; and by the Materials Sciences Division, Office of Basic Energy Sciences of DOE. REFERENCES

1. B. Champagne and R. Angers, “Fabrication of Powders by the Rotating Electrode Process”, Int. J. Powder Metall. Powder Technol., Vol 16 (No. 4), 1980, pp. 359-367.

2. S. Miller and P. Roberts, “Rotating Electrode Process”, Powder Metal technologies and Applications,

ASM handbook-Volume 7, S. Lampman et. al. eds., ASM International, Materials Park OH, 1998, pp. 97-101. 3. B. Champagne, and R. Angers, “Size Distribution of Powders Atomized by the Rotating Electrode Process”, Modern Development in Powder Metallurgy, Proceedings of the 1980 International Powder Metallurgy Conference, Washington, DC, Hausner, H., Antes, H., and Smith G., eds., Metal Powder

Industries Federation, Princeton NJ, Vol 12, 1981, pp. 83-104. 4. Gschneidner, K.A., Jr., Pecharsky, A.O. and Pecharsky, V.K., “Ductile, High Heat Capacity,

Magnetic Regenerator Alloys for the 10 to 80 K Temperature Range”, paper presented at 11 th International Cryocooler Conference, June 20-22, 2000, Keystone, Colorado.

5. Kashani, A., Helvensteijn, B.P.M., Gschneidner, K.A., Jr., Pecharsky, A.O. and Pecharsky, V.K.,

“New Regenerator Materials for Use in Pulse Tube Coolers”, paper presented at 11th International Cryocooler Conference, June 20-22, 2000, Keystone, Colorado.

6. Zimm, C., Jastrab, A., Sternberg, A., Pecharsky, V., Gschneidner, K., Jr., Osborne, M. and Anderson,

I., “Description and Performance of a Near-room Temperature Magnetic Refrigerator”, Adv. Cryog. Engin. vol. 43 (1998), pp. 1759-1766.

7. van Zytveld, J., “Liquid Metals and Alloys” in Handbook on the Physics and Chemistry of Rare

Earth, Gschneidner, K.A., Jr. and Eyring, L., eds., Elsevier Science Publishers B.V., Amsterdam (1989) vol. 12, chap. 85. pp. 357-407.

Magnetothermal Properties of Polycrystalline Gd2ln M. I. Ilyn, A. M. Tishin Physics Department M. V. Lomonosov Moscow State University Moscow, Russia 119899

K. A. Gschneidner, Jr.1,2, V. K. Pecharsky,1,2 A. O. Pecharsky1 1 Ames Laboratory and Department of Materials Science and Engineering Iowa State University Ames, IA, USA 50011-3020

ABSTRACT

The magnetic, thermodynamic and magnetocaloric properties of Gd2In have been investigated as a part of a continuing study of novel lanthanide-based materials in order to improve our knowledge about the nature of magnetic refrigerant materials and to find highperformance alloys suitable for magnetic cooling in various temperature ranges. The dc magnetization, ac and dc magnetic susceptibility, and heat capacity were measured from ~5 to ~350 K in magnetic fields varying from 0 to 100 kOe. Gd2In belongs to the hexagonal Ni2Intype structure. It orders ferromagnetically at ~191 K and then antiferromagnetically at ~91 K. The antiferromagnetic phase is easily transformed to the ferromagnetic state under the influence of a magnetic field. The critical magnetic field varies from ~8.4 kOe at 5 K to zero at 91 K. The magnetocaloric effect, as determined from magnetization data is in good agreement with that calculated from the heat capacity results. The maximum adiabatic temperature rise, as determined from the heat capacity measurements, is 7.4 K at 210 K for a magnetic field change of 100 kOe. INTRODUCTION

Experimental studies of the magnetocaloric properties of materials with unusual magnetic

structures can provide better understanding of the physics of the magnetocaloric effect (MCE). To date, magnetothermal properties have been investigated in a limited number of materials with

certain types of magnetic and structural phase transformations. The physics of the MCE occurring in the vicinity of magnetic order-order phase transitions in rare earth intermetallic compounds is still lacking sufficient experimental and theoretical results. The unusual magnetic properties of the Gd2In intermetallic compound were the reason for carrying out this

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experimental study. Gamari-Seale et al.1, Jee et al.2 and McAlister3 have reported the magnetization and electrical resistivity measurements and found that the compound Gd2In orders ferromagnetically below 190 K and then antiferromagnetically below 99.5 K. It undergoes a metamagnetic phase transition from an antiferromagnet to a ferromagnet in high magnetic fields between ~4.5 and ~99 K. Taking into account what is known about the magnetic structures of the compounds containing other lanthanide elements, which also have hexagonal crystal structures, McAlister3 suggested that the low magnetized state (below 99.5 K) is a spiral antiferromagnetic structure along the c axis. Furthermore, Jee et al.2 considered that there is probably a helical ferromagnetic phase between 99.5 K and 190 K which can be transformed into a simple ferromagnetic state by the application of an external magnetic field. Detailed measurements of the ac magnetic susceptibility in the vicinity of the low temperature phase transition have been carried out by Cowen et al.4 and Saran et al.5 Gamari-Seale et al.1 calculated the paramagnetic Curie temperature and effective magnetic moment assuming Curie-Weiss behavior of Gd2In, and obtained anomalously large value of the effective magnetic moment per gadolinium atom. Based on their magnetic data Szade et al.6,7 suggested that Gd2In does not obey Curie-Weiss law over a broad temperature range above the upper magnetic orderdisorder phase transition. The large value of in this compound and, correspondently, a potentially large MCE, make Gd2In an interesting compound for an experimental study of its magnetothermal properties. EXPERIMENTAL

Polycrystalline Gd2In sample was prepared by arc melting of pure Gd (99.95 wt.% pure) and In (99.99 wt.% pure) in an argon atmosphere. The gadolinium was prepared by the Materials Preparation Center of the Ames Laboratory and the indium was purchased from Johnson Matthey. The sample was re-melted six times with the button being turned over each time to insure homogeneity. No heat treatment was carried out and all measurements were performed using the as-prepared alloy. The x-ray powder diffraction study indicated no impurity phases and confirmed the hexagonal Ni2In-type crystal structure. The space group is P63/mmc, and the lattice parameters are: a = 5.442 Å and c = 6.767 Å, which is in excellent agreement with earlier data. The dc magnetization, and dc and ac magnetic susceptibilities were measured from 5 to 320 K in magnetic fields up to 50 kOe using LakeShore ac/dc susceptometer. The heat capacity was measured using an adiabatic heat pulse calorimeter8 from 3.5 to 350 K in magnetic fields from 0 to 100 kOe. RESULTS AND DISCUSSION

Temperature dependencies of the dc magnetization measured in different applied magnetic fields are shown in Fig 1. As one can see, there are two magnetic phase transitions in low magnetic fields (2 and 5 kOe). At high temperature (above ~280 K), the low magnetized phase is paramagnetic. The inset in Fig 1 represents a plot of inverse dc magnetic susceptibility versus

temperature. It is clearly seen that above ~280 K it behaves linearly. The effective magnetic moment is per gadolinium atom and the paramagnetic Curie temperature is 230.4 K, assuming Curie-Weiss behavior between 280 and 320 K. The value of the effective magnetic moment is distinctly larger than 7.94 µ B theoretically predicted for a free trivalent Gd ion, but it is somewhat lower than that reported by Gamari-Seale et al1. The ordered magnetic moment calculated from the saturated magnetization at 5 K after extrapolation to H = 0 is 7.23 µ B per Gd atom, which is also slightly higher than the theoretically expected 7.0 µ B. During cooling, the magnetization rapidly increased in the vicinity of 200 K and taking into account the

inflection point on the zero magnetic field heat capacity versus temperature (see below), we conclude that Gd2In orders ferromagnetically at ~194 K.

In the low temperature region

(T 10 years) reliability. Contamination continues to remain a primary threat to space cryocooler’s sustained reliability. This paper focuses on the two primary sources of contamination, free gas contaminants, (mixing freely with the helium gas) and bound gas contaminants (such as electrically polarized molecules that stick to metal surfaces and organic vapors evolving over time from

adhesives). Some of the resultant cryocooler degradation scenarios resulting from contamination buildup are discussed. The paper identifies five general sources of contamination, but specifically stresses concern with: Source gas contamination, trapped gas or virtual leaks, and

internal cooler contamination generation. Methods are identified to combat each of these sources. It further cites the need to filter the source gas. The second portion of the paper deals with Raytheon’s experience with contamination, particularly focusing on their SSC #2 cryocooler. After a initial X-ray analysis by Aerospace and a subsequent gas analysis by Pernicka determined that the cooler’s gas was contaminated,

a contamination conference was held at Raytheon which brought in some of the best contamination experts from government and industry to recommend more comprehensive contamination procedures. These are listed in Table 3 of this report. The report then cites what additional measures Raytheon implemented to minimize/prevent any future cryocooler contamination problems with their cryocoolers. Finally, the paper identifies the excellent results Raytheon has had over the last several years with the implementation of their new, more comprehensive procedures.

REFERENCES There are a limited number of reference documents available in this area. Two recommended papers are: 1. Hall, J. L., and Ross, R. G., Jr., “Gas contamination Effects on Pulse Tube Performance,” Cryocool-

ers 10, Plenum Publishing Corp., New York (1999), pp. 343-350. 2. Hall, J.L., Ross, R.G., Jr. and Le, A.K., “A Contaminant Ice Visualization Experiment in a Glass Pulse Tube,” Adv. in Cryogenic Engineering, Plenum Publishing Corp., New York ( 2000) (In Press).

Cryocooler Contamination Study: Temperature Dependence of Outgassing S.W.K. Yuan and D.T. Kuo BAE Systems, Cryogenic Products Sylmar,CA 91342

ABSTRACT

With the advance in technology, the life of the tactical cryocooler has extended way beyond the conventional 4,000 hours milestone1. And with the life extension, the contamination control of the manufacturing process has to be revisited, or cooler life may fall short of expectation, because outgassing is a function of time. This means that cooler components need to be baked out at a higher temperature or for a longer period of time. This in turn requires the knowledge of outgassing

rate as a function of temperature. Contamination study of foreign gases and liquids in a cryocooler was discussed elsewhere2. Outgassing curves of water, alcohol, and acetone at 71°C were presented in the above study. This is a follow-up study emphasizing the temperature dependence of outgassing. INTRODUCTION

As the mean time to failure (MTTF) of cryocoolers approaches and surpasses 10,000 hours, contamination becomes a crucial factor that impacts cooler’s life. Cooler manufacturers are faced with the task of further reducing the beginning-of-life contaminant levels so that the end-of-life levels will still lead to performance that meet the specification. This translates to longer bake-out time at elevated temperatures. To shorten the process time, bake-out at a higher temperature is called for. This requires knowledge of outgassing as a function of temperature. Outgassing properties of various gaseous and liquid contaminants in cryocoolers have been investigated in Ref. 2. Due to the lack of outgassing data as a function of temperature, a linear function was adopted, assuming desorption of gases from a monolayer3. Further testings conducted at BAE Systems indicated that the temperature dependence of outgassing is nonlinear. The test setup and procedure are described in Ref. 2, except that the test chamber temperature was varied between 23°C and 100°C. Figures 1 to 3 show the concentration of acetone, alcohol and water as a function of temperature respectively. As one can see the data show an exponential dependence of concentration as a function of temperature, with much higher outgassing rates at elevated ambient temperatures.

The high outgassing rate at elevated temperature was first attributed to the possibility of phase transition. As the boiling points of the liquids are exceeded, vaporization of the liquids may have resulted in the high contents of concentrations being measured. Further testing on the outgassing property of air shows that it exhibits similar exponential behavior. This confirms that the expoCryocoolers 11, edited by R.G. Ross, Jr.


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CRYOCOOLER RELIABILITY INVESTIGATIONS AND ANALYSES

Figure 1. Concentration of acetone vs. temperature.

Figure 2. Concentration of alcohol vs. temperature.

Figure 3. Concentration of water vs. temperature.

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Figure 4. Concentration of air vs. temperature.

nential outgassing property as a function of temperature is intrinsic to all gases and liquid tested. Using a least squares curve approach, the above data can be fitted with the following equation. (1)

where m is the concentration, T is the ambient temperature, and are constants. Knowing the temperature dependence of the outgassing property, one can then combine with the time dependence function reported in Reference 1 to arrive at the following equation. (2) Where is a function of the initial contaminant level in the system (i.e., the larger the value of the higher the contamination level), B is the temperature dependence constant found in Eq. (1), and

A is the time dependent constant found in Reference 2. The values of these constants are listed in

Table 1. Equation (2) is plotted in Figs 5, 6, and 7 for acetone, alcohol and water respectively at three temperatures, 100°C, 71°C, and 23°C. Experimental data at 71°C are also included. Figures 5, 6, and 7 are extremely useful in devising the bake-out process. The same level of bakeout can be achieved with a much shorter time at an elevated temperature. For example, to reach the same concentration level, the bake-out time for alcohol and water at 100°C is only one tenth that of

71°C. As for acetone, the bake-out time at 100°C is only a quarter that of 71°C.

As the life of cryocoolers extends beyond 10,000 hours3, the tolerance of contaminants in the

cooler becomes smaller. Detection of small level of contaminants at the beginning-of-life can often be difficult especially at relatively low temperatures. The current procedure of BAE’s gas chromatograph (GC) sampling of the working gas calls for two hours of bake-out time at 71°C before

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Figure 5. Acetone concentration as a function of time.

Figure 6. Alcohol concentration as a function of time.

Figure 7. Water concentration as a function of time.

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taking data. By elevating the temperature to above 71C, one can increase the outgassing rate (thus resulting in a more accurate reading of the concentration) and shorten the process time. Although it seems feasible to bake-out coolers at the highest temperature the properties of the cooler materials would allow, precautions must be taken not to expose the motor magnets to temperatures higher than the manufacturer’s specification, which may result in irreversible demagnetization of the magnets. With Eq. (2), one can proceed to calculate the outgassing rate (dm/dt) by taking the derivative of the equation. (3)

The outgassing rates of acetone, alcohol and water at various temperatures are plotted in Figs. 8, 9, and 10, respectively. As mentioned before, outgassing at high ambient is far more effective, with a steep rise between 71°C and 100°C. The outgassing rate is about three times higher in 100C compared to 71C for acetone, and about five times higher for alcohol and water.

Figure 8. Outgassing rate of Acetone.

Figure 9. Outgassing rate of alcohol.

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Figure 10. Outgassing rate of water.

CONCLUSIONS

The temperature dependence of outgassing has been studied in this paper. Outgassing curves of acetone, alcohol and water are presented as a function of temperature. It was found that the outgassing of the acetone, alcohol, water and air, is an exponential function of the ambient temperature. Results of this study are very useful for defining bake-out processes of cooler manufacturing. It can also be used to cut down the process time by baking out components and subassemblies at high ambient temperatures.

REFERENCES 1.

Kuo, D.T., and Yuan, S.W.K, First Order Life Estimation and Its Correlation with Experimental Data, parallel paper in this conference.

2.

Roth.,A., Vacuum Technology, North-Holland (1989), pp. 187.

3.

Yuan, S.W.K., Kuo, D.T., Cryocooler Contamination Study, to be published in Proc. of Advances in Cryogenic Engineering, Vol. 45, 1999.

BAE's Life Test Results on Various Linear Coolers and Their Correlation with a First Order Life Estimation Method D.T. Kuo, T.D. Lody and S.W.K. Yuan

BAE Systems, Cryogenic Products Sylmar, CA 91342

ABSTRACT Life test results of various models of BAE Stirling coolers are presented in this paper together with a first order life approximation model. A cryocooler life estimation method based on the Watt-Hour approach has been developed

elsewhere1. According to this method, the total energy of a cryocooler (i.e., the product of mean input power and total operating time) is conserved. From actual life test data of input power rise as a function of time, the energy of the cooler in Watt-Hour can be calculated by integrating the life test curve. With this knowledge and the specification, one can proceed to estimate life. The biggest disadvantage of this method is that it requires prior knowledge of the life test data before life estimation can be performed. In the present paper, a first order approach is used to estimate the rise in input power as a function of time, which can then be used in life estimation. INTRODUCTION

BAE Systems has conducted life test on various linear motor coolers, including B512C (Ref. 2), B602C (Ref. 3) and B1000E coolers. Conditions of the life tests are summarized in Table 1. Life test data of three different models of coolers in input power vs. time can be found in Figures la, 2a, and 3a. Life test data of cooldown time vs. time are plotted in Figures 1b, 2b, and 3b. And life test data of minimum refrigeration vs. time are presented in Figures 1c, 2c, and 3c, respectively.

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Figure 1a. Input Power of B512C cooler vs. time

Figure 1b. Cooldown time of the B512C cooler vs. time.

Figure 1c. Minimum refrigeration of the B512C cooler vs. time.

BAE’s LIFE TEST RESULTS ON VARIOUS LINEAR COOLERS

Figure 2a. Input power of the B602C cooler vs. time.

Figure 2b. Cooldown time of the B602C cooler vs. time.

Figure 2c. Minimum refrigeration of the B602C cooler vs. time.

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Figure 3a. Input power of B1000E cooler vs. time.

Figure 3b. Cooldown time of B1000E cooler vs. time.

Figure 3c. Minimum refrigeration of B1000E vs. time.


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The life test of the BAE B602C (0.6 W linear) cooler was terminated at 10,008 hours when it

failed the specification. The failure was caused by particulate contamination which blocked the flow in the regenerator. After installing a new regenerator, the performance of the cooler was restored. This proved that the compressor was not the root cause for failure. The life tests of both

the B512C (0.5 W linear) coolers and the B1000E (1 W linear) cooler are still in progress, with the MTTF of the B512C cooler and the B1000E cooler exceeding 9,000 hours and 5,000 hours respectively. Cooler Life Estimation and Correlation with Experimental Data

Cooler life estimation using a Watt-Hour approach was evaluated in Ref. 1. The method assumes that the total watt-hours of a cooler are conserved, i.e., running the cooler at low power

will extend its life and vice versa. Under normal conditions, cooler life is a function of compressor piston wear. As the clearance gap grows (which increases the blow-by losses) due to wear, the driver needs to drive the piston harder to make up for the lost performance. This in turn increases the power (see Fig. 4a). When the power exceeds the user’s specification, the end-of-life of the cooler has been reached. Given the experimental data of the input power increase as a function of time (power curve, Fig. 4a), one can then calculate the total watt-hours of the cooler by integrating the power curve to calculate the total area underneath it. For the example in Fig. 4a, an integration from the initial power of 8 W to the specification of 14 W, gives us the total watt-hours of the cooler that is then entered into Fig. 4b as data point a. The procedure is repeated for various initial powers to come up with Fig. 4b. Of course, if one starts with an initial power of 14 W, the cooler would have essentially no life as indicated in point b of Fig. 4b. A major shortcoming of this method is that it takes extensive life test data to predict life. Moreover, due to the variation in performance from cooler to cooler, data taken from one cooler

may not be applicable to others. This means that a large number of coolers need to be tested before the life of an average cooler can be determined. In this paper, a simple first order model is proposed to estimate the rise in input power as a function of time (power curve). The slope of the power curve (Fig. 4a) is assumed to be proportional to the heat load, the ambient temperature, and the charge pressure, and inversely proportional to the cold tip temperature and the bearing area of the piston seal. The effect of heat load and cold tip temperature on life can be deduced from Figs. 5 and 6. Heat load (cooling capacity) versus input power is plotted in Fig. 5, and heat load versus cold tip temperature is plotted in Fig. 6, for the B1000E cooler. Heat load is almost a linear function of the input power within the range of interest. Large heat loads require high input powers which reduce

the cooler life. The linear dependence of heat load on cold tip temperature (in Fig. 6) suggests that the cold tip temperature is also a linear function of life. With the same input power, the heat load a

Figure 4. Watt-Hour life estimation.

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Figure 5. Heat load vs. input power at 78K.

Figure 7. Input power vs. ambient temperature.

Figure 6. Heat load vs. cold tip temperature.

Figure 8. Input power vs. piston wear.

cooler can alleviate is less at low cold tip temperatures. This means that for the same heat load, one has to apply a higher input power at a low cold tip temperature, which shortens the life of the cooler. In Fig. 7, the influence of ambient temperature on input power is depicted. The data were taken on a B512C cooler at various ambient temperatures at 78 K and with 300 mW heat load. The data appear to be quite linear between 0 to 60°C, with a much sharper rise at temperatures above 60°C. The effect of surface area on life can be obtained from Figure 8, which shows the predicted performance of a cooler (by a second order cooler simulation model) versus compressor piston gap. Since the piston wear is inversely proportional to the piston surface area, and the piston gap is a measure of the piston wear, Fig. 8 implies an inverse linear function between cooler life and bear-

ing surface of the piston. With the above information, one can come up with a simple first order model, by assuming that the slope of the power curve (Fig. 4a) in W/hr is (1) where P is the charge pressure in bars, Q is the heat load applied to the cooler in watts, TH is the and is the cold tip temperature in K. Cont

ambient temperature in °C, A is the piston area is a constant that equals to 5.475E-5.

The slope of the power curve (Fig. 4a) can be calculated by Eq. (1) with the parameters listed in Table 1. The predictions are then compared to the life test data of the three coolers in Table 2.

The model gives a good prediction on the slope of the power curve for both the B512C and the B602C coolers despite a wide range of differences in cooler parameters (Table 1). The correlation

of the first order model with the life test data of the B1000E cooler is not possible due to a decrease in input power (negative power curve slope) as the cooler wears in during the life test. More run


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time is needed in order to validate the model. Equation (1) can be applied to other Stirling coolers, for the trends described in Figs. 5 to 8 are generic to most coolers. In order to apply Eq. (1), one must have some life test data as depicted in Fig. 4a to obtain the constant (Cont) in Eq. (1), With this information, one can proceed to estimate life of different models of coolers of the same design, or coolers of the same model operated at different conditions. As depicted in Figs. 1 to 3, some of the parameters to be monitored during a life test include, input power, cooldown time and minimum refrigeration. Failure in meeting the specification in any of these three criteria constitutes a failure of the cooler. The life estimation discussed in this paper can be applied to all three criteria mentioned above. One simply measures the slopes of the life test data in W/hr (rise in input power vs. time in Figs. 1a, 2a and 3a), minutes/hour (increase in cooldown time vs. time in Figs. 1b, 2b and 3b) and mW/hr (decrease in minimum refrigeration vs.

time in Figs. 1c, 2c and 3c). To estimate the life of the same build of coolers under another set of operating conditions, simply apply Eq. (1). The slopes of the life curves are directly proportional to the ambient temperature, heat load, charge pressure, and inversely proportional to piston area and coldtip temperature. For example, if the ambient temperature is doubled, one would expect the

slopes to be doubled and life of the cooler halved, and if the heat load is halved, the slopes are halved, and cooler life doubled, etc. Precautions must be taken in using Eq. (1), not to exceed the range that this simple approach is intended for. For example, a cooler without any heat load applied will not have infinite life or

zero slope (for the power curve). Also, the equation is only valid for ambient temperatures above 0°C. Generally speaking, the room temperature data is a worst case estimation for life at sub-zero (°C) temperatures. CONCLUSIONS Life test results of BAE’s B512C, B602C and B1000E coolers were presented. The B602C cooler

exceeded 10,000 hours of life test, with the life test of both the B512C (> 9,000 hours) and the B1000E (> 5,400 hours) coolers still in progress. A simple first order life test estimation is suggested, which gave good correlation to the life test data of the B512C and the B602C coolers. More data are needed to validate the model against the B1000E cooler. The effect of charge pressure on life should be further studied. The life estimation method proposed in this paper can be used to predict cooler life as limited by the constraints of input power, cooldown time, and minimum refrigeration.

REFERENCES 1. Kuo. D.T., Loc, A.S., Lody, T.D., and Yuan S.W.K., “Cryocooler Life Estimation and It’s Correlation with Experimental Data,” Advances of Cryogenic Engineering, Plenum Press, NY, vol. 45 (2000). 2 Yuan, S.W.K., Kuo, D.T., and Loc, A.S., “Qualification of the BEI B512 Cooler, Part 1 -

Environmental Tests,” Cryocoolers 10, Plenum Press, NY, pp. 105-110. 3. Yuan, S.W.K., Kuo, D.T., Loc, A.S., and Lody, T.D., “Performance and Qualification of BEI’S 600 mW Linear Motor Cooler,” Advances of Cryogenic Engineering, Plenum Press, NY, vol. 45 (2000).

Initial Observations from the Disassembly and Inspection of the TRW 3503 and Creare SSRB B. J. Tomlinson, Jr. and C. H. Yoneshige

Air Force Research Laboratory Kirtland AFB, NM, USA 87117-5776 M. L. Martin

Dynacs Engineering Albuquerque, NM, USA 87106-4266

ABSTRACT

The importance of long term endurance evaluation data on cryocooler technology cannot be overemphasized. As useful and important as operational data is failure analysis data and the lessons learned from those failures. During fiscal year 1999, the TRW 3503 pulse tube and the Creare Single Stage Reverse Brayton (SSRB) stopped operating during endurance evaluation at the Air Force Research Laboratory (AFRL) Cryogenic Cooling Research Facility (CCRF). On 8-9 November 1998, the TRW 3503 pulse tube cryocooler suffered a system overheat as a result of an environmental control electronics failure. The cooler kept operating throughout the overheat and continued to run after the environmental system was stabilized. The TRW 3503 continued to run with only a small performance degradation of approximately +2K over the design point with a heat load of 0.3W until it tripped and shut down on 20 April 1999. The cooler would not restart, so AFRL engineers and technicians worked with TRW to investigate

the cause of the tripping. On 19 December 1998, the Creare SSRB shut down due to an enviromental system fail-safe overheat condition on the compressor. Subsequent attempts to restart the cooler were unsuccessful. The compressor started normally, but the free spinning turboexpander (TE) was not able to lift off of its bearings and start spinning. AFRL personnel worked with Creare to investigate the cause of the TE’s inability to lift up and spin. This paper will discuss the investigation of both cryocoolers’ inability to restart, initial observations in post-endurance

disassembly, and lessons learned from these experiences. INTRODUCTION

The importance of long term endurance evaluation data on cryocooler technology cannot be overemphasized. As useful and important as operation data is failure analysis data and lessons

learned from those failures. Some of the potential contributors to cryocooler unreliability include wear, performance drift, fatigue, material creep, gaseous contamination particulate/compound contamination and clogging, material and workmanship defects, inadequate machining process Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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development, magnetic circuit degradation, assembly errors, material thermal expansion

mismatches, and long-term alignment instability. In the case of the TRW 3503 Pulse Tube Cryocoler, the environmental system failure and resultant overheat condition pushed this technology far beyond any design condition ever forseen by TRW engineers. Even with the overheat, the cryocooler continued to operate, with only a small degragation in thermodynamic performance, for over 5 months and shut down due to a failure to stake a component, which was previously identified by TRW during the construction of the cooler, and but was not fixed by the government. In the case of the Creare Single State

Reverse Brayton (SSRB) cryocooler, the environmental condition that caused the SSRB to shut down was not nearly hot enough to damage the compressor. However, the system (specifically the turboexpander) would not start due to a balancing problem that has been corrected in flight

hardware development at Creare. AFRL teamed up with both TRW and Creare to investigate and document these failures. TRW 3503 PULSE TUBE CRYOCOOLER The TRW 3503 Pulse Tube Cryocooler is a protoflight unit with a design point of 0.3W @ 35K. It went through characterization at the NASA Jet Propulsion Laboratory (JPL) and AFRL, and then began its endurance evaluation at AFRL. It operated for more than 12,000 hours. This cryocooler was developed under the TRW 35K Pulse Tube Program, and is similar in design to the TRW 6020, the NASA TESS and AIRS cryocoolers, and the Multispectral Thermal Imager

(MTI) Space Cryocooler. Environmental System Failure and Cryocooler Overheat On 8-9 November 1998, the TRW 3503 environmental system electronics failed and the cryocooler significantly overheated. An investigation of the environmental control electronics revealed that one of the two IOTech 488/4 Digital to Analog (D/A) converters failed. The D/A converters link the Lab VIEW™ dataport to the environmental system electronics by converting a digital signal from the dataport to a current, which commands the chiller fluid flow vavles to some increment between fully opened and fully closed. When the D/A converter failed, it stopped sending current signals to the valves. The environmental control electronics read this as a zero amp “signal” and commanded the vavles to fully close, depriving the cryocooler of its heat rejection interface. Examination of the data revealed several interesting things about the temperature excursion. The flow valves for the TRW 3503 were commanded to close at ~2:50 AM on 8 November, which caused the heat rejection temperature and case temperature to rise. The heat rejection temperature continued to rise, surpassing 330K (57° C) at 4:48 AM that same morning. The heat rejection temperature and cryocooler case temperature remained significantly above the bakeout temperature for almost 23 hours. The cooler case temperature reached a maximum temperature of 428.6K and the cold tip temperature reached a maximum temperature of ~65K. Figure 1 shows the heat rejection temperature, the cold end temperature, and the compressor strokes of the TRW 3503 over the period of the temperature excursion. During the entire time that the chiller fluid valves were closed, the cooler continued to operate. As the case temperature rose, the stroke shortened (reaching stroke lengths as small as 9mm compared to the nominal stroke of 12 mm), which is why the cooler didn’t trip and shut down. Once the heat rejection temperature and case temperatures reached steady state at ~9:00 AM on 9 November 1998, AFRL engineers ran a comparative load line with the heat rejection temperature set at 300K. This load line was compared to one that was run on 6 November 1998. As a testament to the robustness of this design, the TRW 3503 endured extremely high environmental temperatures for a significant period of time, and was still able to return to nearly the same performance. There was not much change in the average stroke lengths (the average stroke length required to reach design point on 6 November was 11.93 mm and increased to 11.97 mm on 9 November), and the input power required to reach design point only increased by

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4W after the overheat condition. After the comparative load line, the cooler was returned to its nominal conditions of 0.3W @ 35K, with an average stroke length of 11.97 mm. The TRW 3503 continued the endurance evaluation and continued to run with only a slight performance

degradation of ~2K with a 0.3W heat load. System Shutdown and Diagnosis of the Problem On 20 April 1999, at approximately 4:00 AM, the TRW 3503 tripped and shut down. The Heater Interlock System (a fail-safe electronic circuit to prevent the heater from continually running if the cryocooler stops operation) for the cryogenic heater cut power to the load resistor on the cold block, protecting the system from another overheat. Upon arriving at the lab at ~4:20 AM, an AFRL technician tried unsuccessfully to restart the cooler. From 20-23 April 1999, AFRL engineers and technicians worked with TRW to investigate the situation. Tests were done to eliminate the TRW electronics rack, signal amplifiers and capacitance sensor electronics box. Current displacement and oscilloscope traces taken of compressor 2 and its capacitance sensor showed anomalous readings (Figure 2). The problem lay somewhere inside the cooler, and the decision was made to perform a post-endurance evaluation disassembly and inspection. These activities included an analysis of the working gas inside the cooler, transport of the cooler to TRW for disassembly and failure analysis, and return of the hardware to AFRL for post-disassembly stiction tests and determination of final disposition of the hardware. In Figure 2, channel 1 shows the current probe trace for compressor 2, channel 2 shows the drive output of the TRW electronics, channel 3 shows the drive 2 output of the Techron amplifier, and channel 4 shows the compressor 2 capacitance sensor output. Note the distortions in the compressor 2 current and the capacitance sensor traces show an overall distortion and periodic spikes in the current and displacement.

Figure 1. Case temperatures on the TRW 3503.

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Figure 2. Traces from the TRW 3503.

Gas Sampling and Analysis

On 22 May 1999, the TRW 3503 was taken to Pernicka Corporation in Fort Collins, Colorado for a gas analysis. There was a relatively large amount of carbon dioxide seen in the analysis results. This was to be expected if there were rubbing parts inside the cooler. Also of note is the relatively small amount of.organic materials seen in the gas composition. This provides evidence that the epoxies used in the cooler did not significantly out-gas, even at the high temperatures reached during the overheat condition. 3503 Cryocooler Disassembly at TRW

The cryocooler was taken to TRw for disassembly and a failure analysis on 27-28 May 1999. The compressor housings were removed, and there was no evidence that the flexures were rubbing inside the cooler. The cooler was tested for shorts and all of the readings showed very high or infinite resistance. Then each compressor was run at 5-10 at 24 – 44 Hz. The current traces seen were smooth and clean and showed no indication of a short in the cooler. Upon removing the 9-pin connectors, some feathering was visible on the Teflon casings used to insulated the power leads, which caused some concern, The wires were looked at and photographed under a CCD microscope. The feathering was caused by sharp edges in the holes that the wires were fed through. The holes are in a position that doesn’t allow for de-burring or smoothing on the inside surface of the structure. Stiction tests on both compressors showed no current spikes, which would have indicated that contact was occurring inside the cooler. Increasing the stroke length, however, did show anomalous current traces similar to those seen at AFRL.

Figure 3. Capacitance sensor target with the locking nut detached.


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The next step was to electronically disconnect the capacitance sensors and try running the compressors seperately. Compressor 2 started making a ticking noise. Compressor 1 ran with no problems. The technician at TRW suspected that something might have happened to the capacitance sensor itself. The capacitance sensor in compressor 2 was pulled out of its target, which revealed that the locking nut on the capacitance sensor had backed off and was loose. The loose nut was causing the cooler to trip erroneously by hitting the end stop and simulating over-stroking. The target vibrating inside the cooler caused the rattling noise. In this prototype unit, the locking nut had not been staked to the target. Figure 3 shows the capacitance sensor with the locking nut detached. The locking nut was reattached to the capacitance sensor target, and current traces were taken again. The cooler was taken back to AFRL and current/voltage traces were taken with the TRW 3503’s capacitance sensor box hooked up. Again, the master gain was set at ~0.1. The traces taken after the disassembly show no anomalous spikes or readings, proving that the loose locking nut on the capacitance sensor target was indeed what was causing the TRW 3503 to trip erroneously. CREARE SINGLE STAGE REVERSIBLE BRAYTON CRYOCOOLER

The Creare Single-Stage Reverse Brayton (SSRB) cryocooler was developed under sponsorship by the Ballistic Missile Defense Organization (BMDO), monitored by AFRL and managed by NASA Goddard Space Flight Center (GSFC). The Creare SSRB was designed to lift a 5W heat load at 65K, and operate for 10 years on-orbit. The SSRB technology was chosen by NASA to augment the dewar on the Hubble Space Telescope for the NICMOS system. The SSRB logged over 29,000 hours of operation at AFRL.

Environmental System Failure and System Shutdown

On 19 December 1998, the Creare SSRB automatically shut down when its case temperature exceeded the limit set on its temperature safety circuit. This was not an excessive overheat, it only reached the set point of the safety limit. On 21 December 1998, AFRL personnel investigated the cause of the shutdown and found that the circulating pump motor in the chiller failed. This stopped the flow of chilling fluid to the compressor heat rejection interface. Restart Attempts

An attempt to start the cooler was made on 28 December 1998. The plan was to start the cooler, clean the gas (using the installed getter), and continue with the endurance evaluation. The compressor started normally, but the turboexpander would not lift up and spin (the expander is not driven, it is a free spinning device). During this event, several other attempts were made to restart the cooler, but none were successful. During the restart attempts, a failed power supply in the compressor speed proble signal conditioning circuit was discovered. The circuit board was repaired and a new capacitance probe power supply was built. Upon installation of the new components, two more unsuccessful attempts were made to start the SSRB. Mr. Frank Dolan from Creare came to AFRL on 2-4 March 1999 to assist in troubleshooting the turboexpander (TE) and attempt to restart the cooler without dismantling it. During his visit, the cryocooler was heated using the internal load heaters. This adequately heated the expander bearings closest to the load. The compressor was on during this time, but no spin was detected from the TE. The cryocooler was then evacuated for an extended time, and then refilled with dry neon. The compressor was engaged, but again, no spin was detected from the TE. The vacuum chamber bell jar was then removed, and the TE assembly was heated with a heat gun to ~340K. The compressor was on during this process, but no spin was detected from the TE. As a result, the decision was made to take the TE apart and and inspect it to find out why it would not lift up and spin.

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Turboexpander Disassembly at Creare Frank Dolan and John White traveled to AFRL on 7 Jun 1999 to remove the TE from the SSRB and pack it up for return to Creare. The gas was purged from the system and the multilayer insulation was removed just enough to expose the TE. The TE was removed and all ports were covered with clean room tape to keep the system relatively clean.

The TE was returned to Creare, and systematically dismantled in Creare’s class 10,000 clean room. AFRL personnel were present to observe and assist with the analysis. Components of interest were examined with optical microscopes and photographed. There were several discrepancies noted during the disassembly and inspection. Debris was found in several components, and scratches were found on the surface of the brake inlet and on the TE shaft. Some of the scratches appear to be from the balancing process, while some one of the scratches (seen on the area that would have been covered by the labyrinth seal) looked like rubbing might have caused it. An inspection of the labyrinth seal revealed burrs and sharp edges

that had not been smoothed, which may have caused the scratch on the shaft. Glass beads were also found lodged between the turbine blades at the tip of the shaft. The glass beads are from bead blasting that was done after the shaft was Tiodized. Beads were found in 11 of the shaft passages, and one passage had 3 beads in it. The beads were only found on the exhaust side. No beads were found in the brake channels. Creare admitted that the shaft was not inspected upon its return from Tiodizing. When the TE was lifted manually, it would not drop back down under its own weight. Evem

with separate bearings, the shaft wouldn’t drop under its own weight, but would drop after being

lightly pushed manually. The results of the disassembly and inspection of the TE suggest that the reason the TE could

not lift up and spin is that the shaft was rubbing somewhere inside the housing wall, and the resulting friction made it impossible to lift. One hypothesis is that the shaft was balanced before it was sent to be Tiodized, but after being Tiodized, bead blasted and having glass beads lodged in the passages between the turbine blades, it became unbalanced. Thus, upon its return to Creare, it had to be re-balanced, which caused the numerous scratches seen on the shaft. Rebalancing also caused a high spot on the shaft, which decreased the clearance between the shaft and the housing wall. If any of the glass beads fell out during operation of the cooler, the shaft would no longer be balanced. This might cause it to lean in one direction and rub against the housing wall.

LESSONS LEARNED

In order to prevent another cryocooler temperature excursion, two safety measures were implemented at AFRL. First, the AFRL technicians replaced the faulty D/A converter, and rewired the system so that in the event of a similar D/A converter failure, only one cold plate valve per chamber would close. The other one would continue to function normally and remove heat from the cooler.

The second safety measure implemented was the addition of case temperature inputs to the ISACC 24-hour notification system. If the case temperature of a cryocooler exceeds a certain value, an alarm signal will be sent, and the ISACC system will notify AFRL personnel of the

situation. These two experiences have also emphasized the importance of considering maufacturing processes when designing cryocoolers. TRW now stakes all capacitance sensor target locking

nuts. TRW recognized this as a potential problem prior to and during the development of the 3503, but the government decided! not to open up the cooler and stake the component during the original build. In addition, all coolers with the same heritage as the 3503 were already staked. Since building the SSRB, Creare procured a class 10,000 clean room as well as high powered optical microscopes for inspecting components before they are integrated with the rest of the cryocooler, as well as refined assembly procedures for flight cryocoolers.


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The obvious robustness of these technologies is apparent even through significant overheats and foreign particle contamination. The TRW 3503 would have continued running, at a degraded level of performance, if the capacitance sensor locking nut had been staked. The Creare SSRB was an engineering model non-hermetic cooler and was not designed for long life evaluation and had the Creare SSRB had the level of cleaning now evident in current generation processes, the SSRB would still be continuing in its endurance evaluation. However, these post-endurance evaluations have provided cryocooler developers and users with significan insights into the robustness and the potential for long life of cryocooler technology.

REFERENCES 1. Correspondence with William Burt, TRW, April 1999 to June 1999. 2. Dolan, Frank. “Trip Report – Visit to AFRL to Restart the Creare SSRB Engineering Model Turboexpander”, Creare Incorporated Memorandum (1999). 3. Correspondence with Frank Dolan, Creare Inc., December 1998 to June 1999.

Cryogenic Material Properties Database E.D. Marquardt, J.P. Le, and Ray Radebaugh National Institute of Standards and Technology Boulder, CO 80303

ABSTRACT NIST has published at least two references compiling cryogenic material properties. These include the Handbook on Materials for Superconducting Machinery and the LNG Materials & Fluids. Neither has been updated since 1977 and are currently out of print. While there is a great deal of published data on cryogenic material properties, it is often difficult to find and not in a form that is convenient to use. We have begun a new program to collect, compile, and correlate property information for materials used in cryogenics. The initial phase of this program has focused on picking simple models to use for thermal conductivity, thermal expansion, and specific heat. We have broken down the temperature scale into four ranges: a) less than 4 K, b) 4 K to77 K, c) 77 K to 300 K, and d) 300 K to the melting point. Initial materials that we have compiled include oxygen free copper, 6061-T6 aluminum, G-10 fiberglass epoxy, 718 Inconel, Kevlar, niobium titanium (NbTi), beryllium copper, polyamide (nylon), polyimide, 304 stainless steel, Teflon, and Ti-6Al-4V titanium alloy. Correlations are given for each material and property over some of the temperature range. We will continue to add new materials and increase the temperature range. We hope to offer these material properties as subroutines that can be called from your own code or from within commercial software packages. We will also identify where new measurements need to be made to give complete property prediction from 50 mK to the melting point. INTRODUCTION The explosive growth of cryogenics in the early 50’s led to much interest in material properties at low temperatures. Important fundamental theory and measurements of low temperature material properties were performed in the 50’s, 60’s, and 70’s. The results of this large amount of work has become fragmented and dispersed in many different publications, most of which are out of print and difficult to find. Old time engineers often have a file filled with old graphs; young engineers often don’t know how to find this information. Since most of the work was performed before the desktop computer became available, when data can be found, it is published in simple tables or graphically, making the information difficult to accurately determine and use. NIST has begun a program to gather cryogenic material property data and make it available in a form that is useful to engineers. Initially we tried to use models based upon fundamental physics but it soon became apparent that the models could not accurately predict properties over


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CRYOCOOLER INTEGRATION TECHNOLOGIES AND MATERIALS

a large temperature range and over different materials. Our current approach is to choose a few simple types of equations such as polynomial or logarithmic polynomials and determine the coefficients of different materials and properties. This will allow engineers to use the equations to predict material properties in a variety of ways including commercial software packages or their own code. Integrated and average values can easily be determined from the equations. These equations are not meant to provide any physical insight into the property or to provide ‘standard’ values but are for working engineers that require accurate values.

MATERIALS Initial materials that we have compiled include oxygen free copper, 6061-T6 aluminum, G10CR fiberglass epoxy, 718 Inconel, Kevlar 49, niobium titanium (NbTi), beryllium copper, polyamide (nylon), polyimide, 304 stainless steel, Teflon, and Ti-6A1-4V titanium alloy. These were chosen as some of the most common materials used in cryogenic systems in a variety of fields. MATERIAL PROPERTIES

Thermal Conductivity Widely divergent values of thermal conductivity for the same material are often reported in the literature. For comparatively pure materials (like copper), the differences are due mainly to slight material differences that have large effects on transport properties, such as thermal conductivity, at cryogenic temperatures. At 10 K, the thermal conductivity of commercial oxygen free copper for two samples can be different by more then a factor of 20 while the same samples at room temperature would be within 4%. It is also not uncommon for some experimental results to have uncertainties as high as 50%. Part of our program is to critically evaluate the literature to determine the best property values. Data references used to generate predictive equations will be reported. The general form of the equation for thermal conductivity, k, is

where a, b, c, d, e, f, g, h, and i are the fitted coefficients, and T is the temperature. These are common logarithms. While this may seem like an excessive number of terms to use, it was determined that in order to fit the data over the large temperature range, we required a large number of terms. It should also be noted that all the digits provided for the coefficients should be used, any truncation can lead to significant errors. Tables 1A and 1B show the coefficients for a variety of metals and non-metals. Equation 2 is the thermal conductivity for an average sample of oxygen free copper. It should be noted that thermal conductivity for oxygen free copper can

CRYOGENIC MATERIAL PROPERTIES DATABASE

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vary widely depending upon the residual resistivity ratio, RRR, and this equation should be used with caution. The thermal conductivities are displayed graphically in Figure 1.

Specific Heat

The specific heat is the amount of heat energy per unit mass required to cause a unit increase in the temperature of a material, the ratio of the change in energy to the change in temperature. Specific heats are strong functions of temperature, especially below 200 K. Models for specific heat began in the 1871 with Boltzmann and were further refined by Einstein and Debye in the early part of the 20th century. While there are many variations of these first models, they generally only provide accurate results for materials with perfect crystal lattice structures. The

Figure 1. Thermal conductivity of various materials.

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specific heat of many of the engineering materials of interest here is not described well by these simple models. The general form of the equation is the same as Equation 1. Table 2 shows the coefficients for the specific heat. Figure 2 graphically shows the specific heats.

Thermal Expansion From an atomic perspective, thermal expansion is caused by an increase in the average

distance between the atoms. This results from the asymmetric curvature of the potential energy versus interatomic distance. The anisotropy results from the differences in the coulomb attraction and the interatomic repulsive forces.

Different metals and alloys with different heat treatments, grain sizes, or rolling directions introduce only small differences in thermal expansion. Thus, a generalization can be made that literature values for thermal expansion are probably good for a like material to within 5%. This is because the thermal expansion depends explicitly on the nature of the atomic bond, and only those changes that alter a large number of the bonds can affect its value. In general, large

Figure 2. Specific heat of various materials.


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changes in composition (10 to 20%) are necessary to produce significant changes in the thermal expansion (~5%), and different heat treatments or conditions do not produce significant changes

unless phase changes are involved.8 Most of the literature reports the integrated linear thermal expansion as a percent change in length from some original length generally measured at 293 K, Where is the length at some temperature T and is the length at 293 K. While this is a practical way of measuring thermal expansion, the more fundamental property is the coefficient

of linear thermal expansion,

The coefficient of linear thermal expansion is much less reported in the literature. In principal, we can simply take the derivative of the integrated linear thermal expansion that

results in the coefficient of linear thermal expansion. While we have had success with this method over limited temperature ranges, we have not yet determined an equation form for the integrated expansion value that results in a good approximation of coefficient of linear thermal expansion. For the time being, we will report the integrated linear thermal expansion as a change in length and provide the coefficient of linear thermal expansion when it is directly reported in the literature. The general form of the equation for integrated linear thermal expansion is

Tables 3A and 3B provide the coefficients for the various materials while Figure 3 plots the integrated linear thermal expansions. FUTURE PLANS

We plan to continually add new materials, properties, and to expand the useful temperature range of the predictive equations for engineering use. We will report results in the literature but will also update our website on a continual basis. The initial phase of the program was a learning

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Figure 3. Integrated linear thermal expansion of various materials.

experience on how to handle the information in the literature as well as for the development of a standard set of basic equation types used to fit experimental data. By using just a few types of equations, we hope to make the information easier to use. We shall now focus on developing large numbers of equations for a variety of materials and properties. Please check our web site at http://www.boulder.nist.gov/div838/cryogenics.html for updated information.

REFERENCES 1. Berman, R., Foster, E.L., and Rosenberg, H.M., "The Thermal Conductivity of Some Technical Materials at Low Temperature." Britain Journal of Applied Physics, 1955. 6: p. 181-182. 2. Child, G., Ericks, L.J., and Powell, R.L., Thermal Conductivity of Solids at Room Temperatures and Below. 1973, National Bureau of Standards: Boulder, CO. 3. Corruccini, R.J. and Gniewek, J.J., Thermal Expansion of Technical Solids at Low Temperatures. 1961, National Bureau of Standards: Boulder, CO. 4. Cryogenic Division, Handbook on Materials for Superconducting Machinery. Mechanical, thermal, electrical and magnetic properties of structure materials. 1974, National Bureau of Standards: Boulder, CO. 5. Cryogenic Division, LNG Materials and Fluids. 1977, National Bureau of Standards: Boulder, CO. 6. Johnson, V.J., WADD Technical Report. Part II: Properties of Solids. A Compendium of

The Properties of Materials at Low Temperature (phase I). 1960, National Bureau of Standard: Boulder, CO. 7. Powells, R.W., Schawartz, D., and Johnston, H.L., The Thermal Conductivity of Metals and Alloys at Low Temperature. 1951, Ohio State University. 8. Reed, R.P. and Clark, A.F., Materials at Low Temperature. 1983, Boulder, CO: American Society for Metals.


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Rule, D.L., Smith, D.L., and Sparks, L.L., Thermal Conductivity of a Polyimide Film Between 4.2 and 300K, With and Without Alumina Particles as Filler. 1990, National Institute of Standards and Technology: Boulder, CO.

10. Simon, N.J., Drexter, E.S., and Reed, R.P., Properties of Copper Alloys at Cryogenic Temperature. 1992, National Institute of Standards and Technology: Boulder, CO. 11. Touloukian, Y.S., Recommended Values of The Thermophysical Properties of Eight Alloys,

Major Constituents and Their Oxides. 1965, Purdue University. 12. Veres, H.M., Thermal Properties Database for Materials at Cryogenic Temperatures. Vol. 1. 13. Ziegler, W.T., Mullins, J.C., and Hwa, S.C.O., "Specific Heat and Thermal Conductivity of Four Commercial Titanium Alloys from 20-300K," Advances in Cryogenic Engineering Vol. 8, pp. 268-277.

Experimental Results on the Thermal Contact Conductance of Ag-Filled Epoxied Junctions at Cryogenic Temperatures Z. Wang, A. Devpura, and P.E. Phelan Arizona State University, Mechanical & Aerospace Engineering Tempe, AZ 85287-6106 USA

ABSTRACT

The thermal contact conductance across an epoxied copper junction loaded with Ag (silver) particles was investigated at cryogenic temperatures. Thermal contact conductance, or its inverse, thermal contact resistance, consists of two components: thermal contact resistance at the

copper/Ag-particle epoxy interfaces, and the thermal conduction resistance across the Agparticle epoxy slab. The effects of the Ag-particle volume fraction, and the average interface temperature of the epoxied junction are both evaluated. Increasing the Ag-particle fraction increases the conductance above that for a plain epoxied sample, by as much as one order of magnitude for a Ag particle fraction of 30%. A critical Ag particle volume fraction is observed in the measurements, below which value the thermal conductance of the epoxied junction increases only slightly with increasing particle fraction. INTRODUCTION

When heat flows across an interface between two solids in contact, a relatively high thermal resistance is encountered because of the imperfect junction. The imperfect junction is caused by the roughness of the surfaces on a microscopic scale, such that the real contact surface of the

interface is only a few of the scattered contact spots. Usually, the ratio of the real contact area to the nominal contact area is very small. Thermal contact conductance is defined as where q is the heat flux across the interface, the interfacial temperature drop, and the thermal contact resistance (TCR). In many cases, the energy transfer across the interface is of concern. It is often desired to enhance or isolate the heat transfer at the junction. Therefore, the prediction of is very important.1

The thermal conductance between contacting solids varies considerably, depending upon the thermal and physical characteristics of the materials and the conditions of the junction. In order to control we usually consider several parameters as controlling variables, such as the mechanical and thermophysical properties of the materials composing the junction, the surface characteristics of the contacting surfaces, the mechanics of the contact, and the interstitial materials. In the real world, due to the fixed structure of the entire system, relatively speaking it


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is easy to change the interstitial material. The interstitial materials’ thermal conductivity, hardness and thickness, and the values of the corresponding properties of the base materials, dictate the final value in thermal contact conductance Fletcher2 defined the efficiency as the ratio of the logarithms of the conductances with and without the interstitial material,

and the nondimensionalized conductance,

which takes the thickness of the interstitial material

into account,

where t is the thickness of the interstitial layer and is the equivalent gap thickness. The subscript cm refers to the control material and bj represents the bare junction. These two parameters compare the effectiveness of the insert, which is used for the control of thermal conductance. However, they cannot predict from the operating conditions and the material properties of the interstitial layer. Since there are only limited data available for either for bare junctions or with an interstitial material, especially at cryogenic temperatures, a large number of experiments must be performed in order to determine and for different interstitial materials. Here we investigate one particular category of interstitial materials, Ag-particle-filled Stycast epoxy (Type 2850-FT, Catalyst 9), which may be used at cryogenic temperatures where high values are required. The junction is formed by two copper pieces in a flat-plate geometry. EXPERIMENT Figure 1 presents a schematic of the test column, which is utilized here to investigate an Agfilled epoxied junction, but was previously used to investigate for other materials in contact at cryogenic temperatures.3,4,5 It consists of the test samples, and two calibrated heat flux meters which are pressed together and aligned to the centerline of the apparatus. The alignment can be adjusted by the point contacts of the ball bearings. The test specimens are made of two solid

copper cylinders, epoxied together with Ag-particle-filled Stycast cryogenic epoxy. The thermal conductivity of the heat flux meters was previously calibrated against a reference 304SS bar obtained from the National Institute of Standards and Technology. The temperature drops along the heat flux meters and across the sample interface are measured by differential Type E thermocouples. The differential thermocouples are calibrated at a single point

by immersing one junction into liquid nitrogen and the other into an ice bath, and applying the resulting percent deviation from the standard table to all measurements. The silicon diodes

mounted on the copper heat mounts are utilized as the reference temperatures for the differential thermocouples. The thermocouples are inserted into diameter holes drilled to the center lines of the samples and heat flux meters, and which are partially filled with fine copper powder in order to provide reliable thermal contact. Indium foil is inserted at all junctions besides the sample interface in order to reduce extraneous contact resistance. Two thermal shields are placed above and below the heat flux meters in order to reduce the radiation heat transfer. The entire test column is surrounded by a thermal shield to minimize heat gain from the ambient. The outputs of the thermocouples, silicon diodes, and the load cell amplifier are monitored and recorded by a Macintosh Power PC computer, then simultaneously converted and analyzed by a Labview control program. If the temperature drops across the sample interface as well as those along the upper and lower heat flux meters are all within a set criterion for the last 100 data points, the system is considered to be in steady state. Starting with the reference silicon diodes located on the copper heat mounts, each temperature point is calculated sequentially from the

THERMAL CONTACT CONDUCTANCE OF Ag-FILLED EPOXY

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Figure 1. Schematic diagram of the experimental apparatus.

differential thermocouple readings. The temperature drop at the sample interface is overdetermined by the calculation downwards and upwards, which are averaged to determine the final The heat flux at the sample interface is considered as the average of the upper and lower heat flux meters. All measurements are conducted under a vacuum condition of torr on average, which is measured by a cold cathode vacuum gauge. EXPERIMENTAL UNCERTAINTY The total experimental uncertainty of the thermal contact conductance is from the uncertainties in q and respectively. Since the heat flux q is the average of the measurements of the upper and lower heat flux meters, the uncertainty of is calculated from:6

where and are the uncertainties in q from the upper and the lower heat flux meters, which are calculated from:

where and are the uncertainties in the thermal conductivity K of the heat flux meters and the thermocouple readings, and L and the length and the temperature drop along the heat flux meters.

The total uncertainty in thermal contact conductance,

is calculated from:

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where and are the uncertainties in q and The total experimental uncertainty can be divided into two parts: bias errors caused by instrument specifications and precision errors (95%

confidence level) caused by scatter in the data. For each different error, the final value is calculated from the root sum square of these two parts accordingly. Under cryogenic temperatures, the radiative heat gain from the ambient to the test samples is negligible. The maximum radiative heat gain was estimated to be less than 5 mW, while the

normal heat flow along the axial direction of the sample is about 2.5 ~ 3.0 W. Both junctions of the differential thermocouple positioned across the sample interface are mounted in holes located 2.4 mm from the interface. The resulting temperature drop across the copper is estimated as 1.4% of the across the interface and is considered to be negligible.

SAMPLE PREPARATION AND CHARACTERISTICS

The purpose of this investigation is to determine how

of filled epoxied junctions varies

with the particle volume fraction at cryogenic temperatures. The temperature ranges from 40 to

230 K. At these low temperatures, conventional epoxies are prone to cracking due to thermal expansion mismatch. Hence, our experiments are performed with an epoxy that is designed specially for cryogenic use. By analogy with the percolation behavior of electrically conductive composite materials,7 we know that the larger the difference between the thermal conductivities of the filler and the matrix materials, the more liable we are to observe the percolation phenomenon in the experiments. Pure silver has one of the highest thermal conductivities among all the metals, and it is very stable under typical experimental conditions. Therefore, the metal basis silver powder with 3N purity, purchased from Alfa Aesar, was chosen as the filler material. Basically, the particles are not spherical, as they have irregular shapes, but Alfa Aesar provides a size parameter to describe the particles. The size ranges from For the epoxy, there are several suitable cryogenic candidates. By considering, however, that the particle distribution of the samples will be examined by an optical microscope, we select Stycast epoxy, having a dark color, as the matrix material. This provides good contrast, and makes it easy to distinguish the white Ag filler particles from the black epoxy background. The surface treatment method and roughnesses for the copper samples are tabulated in Table 1, where the surface parameters (roughness) and m (mean slope) are defined elsewhere.3 Note that all surface measurements were performed at room temperature. The surface conditions of all


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samples are similar to one other. A certain amount of silver powder and Stycast epoxy are blended together, according to the desired Ag volume fraction

Since the densities of both materials are available,

can be

controlled by controlling the mass of silver powder mixed in with the epoxy according to

where and are the masses of Ag particles and of the epoxy, and the densities of Ag and of the epoxy, and the mass of epoxy removed before mixing, in order to arrive at the desired

The thickness of the epoxy mixture layer is constricted to the thickness of a filler gauge (250 After the experiments, the actual thickness was examined under an SEM (Scanning Electrical Microscope). Table 2 shows the measured thicknesses and slopes, i.e., deviations from a perfectly parallel interface where the slope would be zero degrees, of the four samples. In order to observe the distribution of Ag particles in the epoxy, we examine a cross section of the epoxy under an optical microscope, with 100X magnification. The image is subsequently digitized for analysis, and the minimum and maximum of the x, y coordinates are set as 0.0 and 1.0. Figure 2 shows a local image of the particle distribution, for the sample containing 25% Ag particles. It was desired to gauge the randomness of the Ag particle distribution. Accordingly, ChiSquare tests were performed separately on the x and y position coordinates:8

Figure 2.

A partial image of the cross-section for the 25% Ag fraction sample under an optical microscope (original image with 100X magnification).

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Figure 3.

where

Comparison of

and

among a pure epoxy junction and four Ag-filled epoxied junctions.

are the observed and expected values in the ith interval,

the population

variance, and k the number of intervals. For the test sample having 25% volume fraction of Ag particles, the results show that the Ag particles are randomly distributed over the cross-section. The correlation between the x and y coordinates, r, was also found to be close to zero (less than 0.05), so that the following conclusion can be made:9 the Ag-particles are uniformly distributed in the epoxy, with a confidence level of 95%. Furthermore, since the cross-section examined was chosen randomly, we conclude that the distribution of particles is random throughout each

of the epoxied junctions. EXPERIMENTAL RESULTS AND DISCUSSION Four test specimens were investigated for the of Ag-filled epoxied junctions. As was shown in Table 2, the volume fractions of Ag particles were approximately 10%, 20%, 25% and 30%, respectively. Separate experiments were conducted for these four different Ag volume fractions. Great effort was made to maintain the consistency of all experimental conditions for all

the samples during the measurements, so the difference among the volume fractions can be counted as the only variable to explain the different

results.

Figure 3 shows the experimental of the four samples compared with that of a pure epoxy junction sample.10 The thickness of the pure epoxy junction is close to the average thickness of the Ag-filled ones, and other experimental conditions are the same, so the difference in is primarily due to the Ag particles. AH four different Ag fraction curves show clearly a trend that increases with an increase in the average junction temperature. The rate of increase is larger in the low-temperature range, while in the high-temperature range, the curve becomes more flat. Around 50 K, there is a big leap in thermal contact conductance. For is very close to that for the pure epoxy data. There is a trend, though still not obvious, that the data are a little higher than those of the pure epoxy after a pivot temperature (150 K).


Figure 4.

Effect of varying the Ag-particle fraction on and 70 K).

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(junction temperatures near 100 K, 85 K,

The curve shows that the two sets of data are still close to each other, but the pivot temperature is lower (110 K). At low temperatures (below 70 K), the and pure epoxy curves are close, however, above 70 K there is a pronounced increase in of the sample over that of the pure epoxy. For is greater than that for pure epoxy over the

entire temperature range. The difference between the Ag-filled and pure epoxy increases with increasing while the pivot temperatures fall with increasing The existence of the Agparticle filler in the Stycast epoxy apparently enhances as expected, with the enhancement being the most dramatic for

The error bars of the measurements, as displayed in Fig. 3, are generally small, except in the low-temperature range (less than 50 K). As the average temperature at the interface rises, the error decreases rapidly.

Figure 4 shows the effect of varying the Ag-particle volume fraction on at several values of temperature (100 K, 80 K, and 70 K). The participation of the Ag-particles in the junction changes differently according to the magnitude of For the first three samples 20%, and 25%), there is only a small increase in However, (30%) for all temperatures is special in the figures: there is a trend to rapidly increase The of the junctions includes two parts, the thermal conductance of the epoxy composite, and the two thermal boundary resistances at the interfaces, which add to determine the overall

where is the effective thermal conductivity of the epoxy composite, and is assumed to have the same value at each epoxy/copper interface. The value of is determined, in part, by the samples’ surface conditions. Table 1 shows that the surface roughnesses of all samples are relatively close to one other. We therefore assume that remains unchanged from sample to sample, leaving as the only factor that influences Here, we estimate from the value determined from two pure Stycast epoxied samples, giving at

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The values of effective thermal conductivity can now be extracted from the measured using Eq. (8), and the resulting values are given in Table 3. Two approaches that provide upper and lower bounds to where is the thermal conductivity of the epoxy matrix, are to (i) assume the matrix and filler lie in series (lower bound), and (ii) assume the matrix and filler lie in parallel (upper bound). The corresponding relation for the series approximation is:11

and that for the parallel approximation is:11

For the values of calculated from Eqs. (9) and (10) are shown in Table 3, together with the experimental results. All the experimental data lie between the upper and lower bounds. For the first three fractions (below 30%), the parallel approximation is much higher than the experimental data, while the series approach makes a better estimation. However, for the last fraction (near 30%), the experimental data are substantially larger than the lower bound, and draw closer to the upper bound. The parallel approximation can be interpreted as the extreme percolation case, in which all the Ag particles are in contact and form a path through the slab, and there is no thermal contact resistance between the Ag particles. In reality, however, the thermal contact resistance between the Ag particles is non-negligible, and not all the Ag particles come together to form a high-conductivity path across the slab. Therefore, the measured value of is lower than that given from the parallel approximation.

CONCLUSIONS The thermal contact conductance for four different Ag-particle volume fractions of filled epoxied junctions is investigated at cryogenic temperatures, where the average interface temperature ranges from 40 K to 210 K. Thermal contact conductance increases as the junction temperature increases, with the increase in the low-temperature range being more significant. The results indicate that increasing the Ag-particle volume fraction increases in accord with expectations, for the greater the number of Ag particles, the greater the possibility that thermal shortcuts of Ag particles are formed. The presence of the Ag particles enhances beyond that measured for a plain epoxied junction, with the enhancement for the curve being as much as one order of magnitude.


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ACKNOWLEDGMENTS

All the authors wish to thank Lisa De Bellis for her assistance in the experiment. P.E.P. gratefully acknowledges the support of the National Science Foundation through a CAREER Award (Grant No. CTS-9696003), and matching funds provided by Raytheon. REFERENCES 1. 2. 3.

Madhusudana, C.V., Thermal Contact Conductance, Springer, New York (1996), pp. 77-99. Fletcher, L.S., “Recent Developments in Contact Conductance Heat Transfer,” J. Heat Transfer, vol. 110 (1988), pp. 1059-1069. Phelan, P.E., & Mei, S., “Experimental Results on the Thermal Contact Resistance of G-10CR

Composites at Cryogenic Temperatures,” Paper No. AJTE99/6156, Proc. of the 5th ASME/JSME Joint Thermal Engineering Conference, San Diego, CA, March 15-19 (1999).

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Zhao, L., & Phelan, P.E., “Thermal Contact Conductance Across Filled Polyimide Films at Cryogenic Temperatures,” Cryogenics, vol. 39 (1999), pp. 803-809.

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De Bellis, L., Phelan, P.E., Drake, P., & Kroebig, W., “Measurement of the Thermal Properties of

Epoxied Titanium Contacts at Cryogenic Temperatures,” to appear in Advances in Cryogenic Engineering, vol. 46 (2000).

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Holman, J.P., Experimental Methods for Engineers, McGraw-Hill, Inc., New York (1994), p. 66.

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Joy, T., “Percolation in a Thin Ply of Unidirectional Composite,” J. Composite Materials, vol. 13,

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Banks, J., & Carson, J.S., Discrete-Event System Simulation, 2nd Ed., Prentice-Hall of India, New Delhi (1998), p. 303.

(1979), pp. 72-78.

9. Swartz, C.E., Used Math, State University of NY at Stony Brook, New York (1973), p. 100. 10. De Bellis, L., & Phelan, P.E., “Measurement and Prediction of the Contact Conductance Across Epoxied Copper Contacts at Cryogenic Temperatures,” to be presented at the ASME International

Mechanical Engineering Congress & Exposition, Orlando, Florida, November 5-10 (2000). 11. P. E. Phelan, P.E., & Niemann, R.C., “Effective Thermal Conductivity of a Thin, Randomly Oriented Composite Material,” J. Heat Transfer, vol. 120 (1998), pp. 971-976.

A Fail-Safe Experiment Stand for Cryocooler Characterization C. H. Yoneshige1, N. S. Abhyankar2, J.P. Kallman1, G. W. Lybarger1 and M. L. Martin2 1

Space Vehicles Directorate, Air Force Research Laboratory Kirtland AFB, NM, USA 87117-5776 2

Dynacs Engineering Co Albuquerque, NM, USA 87106-4266

ABSTRACT

Understanding possible failure mechanisms is the first step in the prevention and impact minimization of experiment stand failures in any research facility. As the facility encompasses a wider range of multiplicity of experiments, these mechanisms can become coupled, requiring careful planning and preventative measures. This paper will discuss the possible experiment stand failure modes and safeguards used to minimize the impact that experiment stand failures have on the cryocooler characterization process at the Cryogenic Cooling Research Facility at the Air Force Research Laboratory (AFRL). The facility is equipped with one 6 foot, two 18

inch, two 36 inch and eight 24 inch thermal vacuum chambers along with all of the equipment, specialized tools and instruments required for the short term performance characterization and long term endurance evaluation of cryogenic coolers. The laboratory characterizes cryocoolers based on various thermodynamic cycles, such as the Stirling cycle, pulse tube (Stirling cycle

variant) and reverse Brayton cycle, at various levels of technical maturity. These cryocoolers were developed to satisfy space mission requirements from the United States Air Force (SBIRSLow Program Office), The Ballistic Missile Defense Organization (BMDO), the National Aeronautics and Space Administration (NASA) and the Department of Defense. The prevention of experiment stand failures related to cryocooler characterization will be discussed from different standpoints, including the vacuum system, equipment interfaces, electronics, software programs, communication systems, and electronic remote notification systems. INTRODUCTION

Failures may occur on the one component that was assumed to be “non critical” and was therefore not addressed in the fail-safe set up of the experiment stand. Most potential component failures are accounted for during routine maintenance. However, every so often, a component fails and produces unexpected results. This leads to the reevaluation of the experiment stand setup and the revision of the routine maintenance list. The experiment stand fail-safe set up is Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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critical for 24-hour laboratory operations encompassing multiple experiments with extensive instrumentation.

The Air Force Research Laboratory Cryogenic Cooling Research Facility (CCRF) characterizes unique, space qualifiable cryocoolers for thermodynamic and long term (~10 years) performance in a space like environment. The CCRF was built and maintained under joint sponsorship by the Ballistic Missile Defense Organization (BMDO) and the United States Air Force (USAF). The facility is equipped with two 18”, eight 24”, two 36”, and one 6’ thermal

vacuum chambers. Most cryocoolers are characterized in the thermal vacuum chambers. However, coolers that are not space qualifiable, but have heritage to space cryocooler designs are tested on table-top experiment stands with a vacuum bonnet covering the cold end only. Cryocooler development is supported by BMDO, AFRL, the SBIRS-Low program office and NASA to meet specific mission requirements. The lab has characterized cryocoolers from TRW, Raytheon, Ball Aerospace, Creare, Lockheed Martin, Matra Marconi Space, and other agencies. These cryocoolers are based on different thermodynamic cycles, such as the Stirling cycle, pulse tube (Stirling cycle variant), and reverse-Brayton cycle. Each cryocooler’s

thermodynamic performance envelope is established. Then it is put into an endurance evaluation and run until it meets predetermined failure criteria. Cryocooler characterization involves a series of functional experiments, such as load line trials, temperature stability trials, transient thermal response trials, cool-down curves, and parasitic heat load determination, which explore the cryocooler’s thermodynamic capability/capacity as well as its compatibility with its proposed spacecraft, its mission and its

subsystems. A complete performance envelope is determined for each cryocooler, providing a range of operation points which can be used by space mission planners to find existing cryocoolers suited for their specific mission requirements. The endurance evaluation monitors the cryocooler’s long-term behavior and helps to build confidence in new cryocooler technologies. It also provides insight into failure mechanisms. All of the data collected provides feedback to cryocooler researchers and manufacturers regarding design improvements based on

lessons learned. A typical experiment stand for cryocooler characterization consists of the following subsystems: - Vacuum System - Temperature Rejection System and Interfaces - Electronics - Software All of these subsystems interact while the experiment is running. Failure modes for each subsystem must be addressed both independently and collectively during cryocooler integration. This paper addresses the identification of critical system components and the implementation of safety features to protect experimental hardware in case of a component failure. IDENTIFICATION OF CRITICAL COMPONENTS AND FAIL-SAFE FEATURES

A typical experiment stand used for cryocooler characterization is shown in Figure 1. This experiment stand is set up for a split-Stirling cryocooler and includes the heat rejection interface. Cables, fluid lines, and thermocouple wires are passed through air-tight feedthroughs which are located around the bottom of each vacuum chamber. Remote Notification System While a cryocooler is operational, critical operation parameters are continuously monitored

to make sure they stay within their specified limits. If the limit is crossed, an alarm condition exists, and a signal is sent through a custom-built alarm interface box to the Intelligent System for Automatic Control and Communication (ISACC) remote notification system. The ISACC system can monitor equipment and environmental conditions using 16 analog or digital inputs. It can also be programmed to switch up to 8 digital outputs and four analog outputs in any

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Figure1. A typical cryocooler characterization experiment stand.

combination. If an alarm is detected, ISACC can dial up to 8 phone numbers and communicate in voice or data. One technician or engineer is placed on-call each week and carries a pager, and is the first person called in case of an alarm. If this person cannot be reached, ISACC is programmed to call other lab personnel. Utilization of this system allows cryocooler

characterization experiments to be run unattended 24 hours a day, seven days a week. Specific alarm conditions that have been set will be covered in the following sections.

Vacuum System The vacuum system consists of a roughing (vane) pump, turbo pump, gate valve, and instrumentation feedthroughs. In this section, only the mechanical aspects of the vacuum system will be addressed. The electronics and instrumentation will be covered in a later section. Possible

component failures and their consequences are described, and their fail-safe solutions are presented. Some of the possible component failures include the following: Pump Failures. If the turbo pump fails, and the chamber pressure is lower than the

minimum pressure that the roughing pump can sustain, oil from the roughing pump can back stream into the chamber and contaminate the surfaces. As a result of this risk, a system was set up to close the gate valve in the event that the pressure in the vacuum chamber exceeds a setpoint value. A vacuum loss alarm was also set up and connected to the remote notification

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system. If the roughing pump fails, the turbo pump will fail and cause an increase in the vacuum pressure, again triggering the vacuum loss alarm. Gate Valve Failure. The gate valve is pneumatically operated (positive pressure is required for the gate valve to remain open). If the shop air compressor fails or AC power is lost, the gate valve will close, protecting the vacuum inside the chamber. Heat Rejection System and Interfaces

When a cryocooler is operating in orbit, the heat it generates is normally rejected to deep space through a radiator. In the laboratory heat is rejected to a surface, which is kept at a constant temperature by alternately running a refrigerant through a chiller loop and heating it with a heater element. The optimal heat rejection temperature for each cryocooler is specified in its design specifications. The fluid from the chiller flows through a spiral of copper pipes which lie under the heat rejection surface. Heat is exchanged with the cryocooler by conduction at the heat

rejection surface and then by convection with the chiller fluid, which returns to a constant temperature chiller reservoir. The fluid enters and exits the chamber through special feedthroughs. The amount of fluid flowing through the system is determined by flow valves, which are controlled manually or by an electronic controller. If any part of the heat rejection system fails, the risk of a system overheat, thermal runaway, or loss of vacuum exists. Operation at high temperatures can damage the cryocooler. Some of the

possible component failures include the following: Chiller. The chiller is probably the most sensitive component of the heat rejection system. It can be affected by commercial power outages and has many parts succeptible to failure. The pump motor, refrigeration unit and plumbing lines are all possible failure mechanisms. Precautionary measures were taken to ensure that any chiller component failures would not cause damage to the cryocooler under evaluation. In case of a commercial power failure, the chillers are hooked up to a back up power generator, which is activated within about 30 seconds of the initial power loss. The chillers are programmed to restart once the generators are turned on. Alarms were also set up to notify lab personnel in case of any component failures. If the pump motor fails, there is a direct interface to the ISACC system. If there is a leak in the fluid lines, there is a low-flow alarm, which notifies lab personnel if fluid flow falls below a specified limit. If the leak is inside the vacuum chamber, the chiller fluid will increase the pressure inside the chamber and the vacuum loss alarm will activate. Flow Valves. The flow valves are controlled by computer software which sets the amount that the flow valves are open to maintain the desired flow rate. The valves are set to fail open,

which protects the test article from overheating. Heating Units. Either serpentine heaters or cal rods heat the chiller fluid and maintain a specified heat rejection temperature. The method used for heating the chiller fluid depends on the requirements of each individual cryocooler or cryogenic technology experiment. The only consequence of a heating element failure is the inability to run the experiment at a higher heat rejection temperature. This poses no risk to a cryogenic experiment. Electronics

There are four electronics subsystems used in the experiment stand. One subsystem monitors and controls the environmental system, another consists of instrumentation used to collect data from the experiment, another consists of power supplies use to run the cryocooler and its electronics and a fourth is used to monitor the experiment stand and activate the remote notification system. One last item – probably the most important – is the set of cables used to connect all the electronic subsystems together. Environmental. The environmental system electronics consists of controllers and instrumentation used to maintain a vacuum inside each thermal vacuum chamber and control the


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Figure 2. Vacuum Chamber Environmental Electronics.

heat rejection temperature. Figure 2 is a block diagram showing a typical vacuum chamber environmental set-up. The Tempscan monitors the environmental thermocouples. It is also used as a digital I/O device which controls the vane pump and opens or closes the gate valve through relays, indicated by a letter R in the figure. If the Tempscan fails, the data from the environmental thermocouples

is lost, subsequently the gate valve closes and the vane pump shuts off. There is no threat to the test article if the data from the environmental thermocouples is lost. Closing the gate valve protects the vacuum inside the chamber for a little while. When the vacuum pressure increases enough, the vacuum loss alarm is triggered and lab personnel are notified. The bus converter provides serial communication between the General Purpose Interface Bus (GPIB) and the chiller, Varian and Omegas. The Varian reads the vacuum pressure and updates the environmental software. The omegas control the flow valves. If the bus converter fails, the chillers will continue to run, the Varian will lose the ability to update the vacuum

pressure reading on the software, and the omegas will keep running at the last input setting. None of these consequences poses any threat to the experiment running in the chamber. However, if one of the Omegas fails, the flow valves will close. A zero signal equates to fully closed. This deprives the test article of the chiller fluid that keeps the heat rejection

temperature down. For this reason, thermocouples were placed on the warmest running parts of the cryocooler compressor (and expander, if applicable) casing. The readings from these thermocouples are fed into a case temperature alarm interface box. If the case temperature goes above the alarm limit, the case temperature alarm activates and notifies lab personnel. If the Varian ion gage fails to get emissions, an alarm is activated. If the turbo pump controller fails, the turbo pump stops functioning, the vacuum pressure rises and the vacuum loss alarm is activated.

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Data Acquisition Instrumentation. The two major types of instrumentation used to monitor cryocooler characterization are thermometry instruments and power meters. Thermocouples measure the ambient temperature in the vacuum chamber, and are attached to different parts of the cryocooler and test stand to provide a thermal map of the experiment. Diodes and Platinum Resistance Thermometers (PRTs) measure the temperature of the cold head (and each additional stage of a multi-stage cryocooler). Fluke Data Acquisition Units monitor the thermocouples, and either Lakeshore or Conductus temperature controllers monitor the diodes and PRTs. Valhalla power meters monitor input power (in terms of current and voltage) to the cryocooler. The drive power is wired through the Valhalla. If the Valhalla fails, the input power continues to go through to the compressor. The only loss is the input power data. The same holds true if a Fluke, Lakeshore or Conductus fails. The data losses pose no threat to the cryocooler. All of the data acquisition instruments are on a calibration schedule to make sure that the data is traceable to NIST standards. This calibration cycle also helps find problems with the equipment and keep up on maintenance. Power Supplies. Some cryocoolers require external power supplies. Others have their

power supplies inside their control electronics. If any of the power supplies fail, the cooler stops running. If a cooler stops running or trips, an alarm is activated to alert lab personnel to the situation. Cables. Failures can also occur in cables at their connector junctions. The cables run from the electronics boxes or instruments through the vacuum chamber feed-throughs and to their respective components inside the chamber. The wires are individually soldered, and can come loose if the cables are bent or moved on a regular basis. This can cause intermittent contact,

which is one of the most difficult problems to diagnose. The best way to prevent cable failures is to be extremely careful when building the cables and harnesses. Once the cables are built, they go through an extensive visual and mechanical inspection and bench testing procedure. Software

In terms of the experiment stand, the only software concerned is the environmental control software and the data acquisition software. Both sets of software were written in-house using Lab VIEW™ software. The environmental software monitors and controls the environmental

conditions in each thermal vacuum chamber. The environmental software is not used in every vacuum chamber. In some cases, it is easier to control the vacuum system manually. The data acquisition software monitors and records parameters, which include input power, input voltage and input current and cryocooler cold end, body and heat rejection temperatures. The external heat load is also set through this software. Some cryocoolers come with stroke length monitoring devices, such as LVDT or capacitance sensors. Their responses in terms of voltage are converted into stroke lengths in terms of millimeters. The conversion factors are provided by the cryocooler manufacturers and verified by AFRL engineers. Figure 3 shows a data acquisition front panel for Raytheon’s PSC, a protoflight spacecraft cryocooler. The graphs are for endurance evaluation at its design point of 60 K at 2 watts of load. The software programs, in general, monitor the system for normal operation and for outside of its set limits. The operation limits can be set and adjusted from the software. Based on these limits, alarm conditions are established. If an alarm condition is detected, the system is set up to take data in smaller time intervals, giving lab personnel a better picture of what causes the alarm to activate and the subsequent consequences.


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Figure 3. Data Acquisition Front Panel. CONCLUSIONS

In the course of developing the Cryogenic Cooling Research Facility, many lessons were learned regarding the fail-safe set-up of experiment stands. In an effort to avoid over-alarming the system and wasting time and resources, the consequences of each component failure were considered and alarms were placed where they could look for those consequences. It turned out that a few alarms could cover multiple component failures. One of the greatest lessons learned is the importance of effective communication between the engineers and technicians. With everyone clear on what direction each experiment was

taking, the proper fail-safe mechanisms were put into place, leading to systems that operated longer and provided greater protection for experiments running unattended.

Development and Testing of a Gimbal Thermal Transport System D. Bugby, B. Marland, and C. Stouffer Swales Aerospace Beltsville, MD, USA 20705 B. Tomlinson, T. Davis

Air Force Research Laboratory Kirtland AFB, NM, USA 87119

ABSTRACT

This paper presents development details of a two-phase thermal transport system for carrying cryocooler waste heat across a 2-axis gimbal with low line-induced torque. Applications for this system include space-based remote sensing spacecraft with gimbaled cryogenic optics and/or infrared sensors. The described system will use standard loop heat pipe (LHP) technology and small diameter stainless steel bellows tubing to transport 50-200 W of waste heat over a distance of a few meters while still meeting strict gimbal flexibility requirements. The gimbal will be motorized in each axis to study whether rapid gimbal movement affects thermal transport system operation. Heaters will simulate cryocooler waste heat. Preliminary results from a

pressurized, multi-line, non-motorized gimbal mock-up indicated the line-induced torque could be acceptably low if the routing is appropriately managed. The paper will describe the design and development effort. Test results were not available at the time the paper was written. INTRODUCTION

Cryocooled infrared payloads are being considered increasingly by the DoD and NASA for space-based missions that require advanced imaging and tracking capabilities. In many instances, these systems operate best when the optics/sensors can be independently slewed relative to the spacecraft. Thus, a gimbaled optical/sensor system is the preferred approach, but the problem becomes how to reject the heat generated by on-gimbal cryocoolers to space. Typically, ongimbal radiators cannot be used due to their poor radiative efficiency, large weight/volume, and adverse impact on system performance. And, as cryogenic cooling loads increase in the future, this thermal management problem will become even more severe. In addition, the problem of

transporting heat across an isolated interface is also a concern for infrared detectors and optical benches that need to be structurally isolated. There is also a similar problem when attempting to use a deployable/steerable radiator to reduce the size of a spacecraft launch vehicle and meet a reduced envelope. The solution to this long-standing thermal problem is the subject of this paper.

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The solution is to utilize appropriately routed stainless steel bellows tubing and an ambient loop heat pipe (LHP) two-phase heat transport system to continuously move cryocooler waste heat across the gimbal to a spacecraft-mounted radiator. Although typical LHP systems are completely passive with no moving parts, the challenge is to manage the line routing so that the gimbaled sensor/optical system can still meet its requirements for 2-axis motion, low torque, low jitter, high reliability, and long lifetime. The principal technological uncertainty is whether gimbal motion during LHP operation has any unknown dynamic and potentially deleterious effect on the system's ability to meet mission requirements. Figure 1 illustrates the concept.

To meet the technical need outlined above, this paper presents the development details of an AFRL funded program to design, fabricate, and test a two-phase thermal transport system for carrying ambient waste heat across a 2-axis gimbal. This program, which is part of a broader AFRL initiative known as CRYOBUS (see Bugby1), has been given the name GATTS (Gimbal Ambient Thermal Transport System). The paper is organized as follows. First, technical background information is provided on infrared sensor pointing methods, LHP technology options, and flex line routing options. Next, the overall program is described including the GATTS objectives, requirements/specifications, design calculations/trades, design approach, and test plans. Lastly, the conclusions of the program and future plans (at the time this paper was

written in June 2000) are outlined and discussed. BACKGROUND The following presents background information on infrared sensor/optical system pointing methods, LHP technology options, and flex line technology.

Pointing Methods

There are four possible approaches for pointing an infrared sensor/optical system at a target. Arranged from least to most agile, the advantages/disadvantages of each are described below. Spacecraft Pointing. In this option, the infrared sensor system is contained entirely on the spacecraft and the spacecraft is slewed toward the target. The advantages of this pointing method are simplicity (no additional optical components are required), narrow field of view (so the optics can be well baffled and stray light will be minimized), and spacecraft compactness. The disadvantages are poor pointing agility (large momentum wheels), high power, degraded communication/power during pointing (re-pointing required for high gain antennas and solar arrays), and degraded space-view during pointing (important if cryogenic, passively-cooled detectors are used).

Figure 1. Notional Representation of the Across-Gimbal Thermal Transport System

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Instrument Pointing. In this option, the entire instrument is contained within a canister on a 2-axis gimbal. The advantages of this pointing method are pointing agility (the mass/momentum disturbance is less than that of the spacecraft pointing option), uninterrupted communication/power during pointing (the orientation of antennas/solar arrays/radiation coolers is determined by the spacecraft attitude), simplicity (no additional optical components are required), and narrow field of view (so the optics can be well baffled and stray light will be minimized). The disadvantages are mechanical complexity (a large canister on a 2-axis gimbal is needed), and electrical/thermal complexity (power, data, and thermal transport lines must be routed from the telescope across the gimbal to the spacecraft). Telescope Pointing. In this option, a portion of the optical system is included on a 2-axis gimbal, and the rest of the system is on the spacecraft. The advantages of this approach are

pointing agility (the mass/momentum disturbance is less than that of the instrument pointing option), a narrow field of view (so the optics can be well baffled and stray light will be minimized), and system simplicity (power, cryogenic lines, and focal plane arrays (FPAs) are located on the spacecraft). The disadvantages are optical complexity (the telescope's optical path must be split between the canister and the spacecraft with at least one additional fold mirror) and image rotation. Scan Mirror Pointing. In this option, the instrument is on the spacecraft and a scan mirror is on a 2-axis gimbal. The optical system has a wide field of view and the scanning mirror is used to acquire and track targets. The advantages here are pointing agility (the mass/momentum disturbance is less than that of the telescope pointing option), optical simplicity (all curved telescope optics and focal planes are on the spacecraft body), and electrical/thermal simplicity (power, cryogenic lines, and FPAs can be located on the spacecraft body). The disadvantages are mechanical complexity (the system uses scan mirror on a two-axis gimbal), image rotation, and

stray light (due to the wider field of view). Conclusion. For cryogenic infrared sensor applications where stray light must be minimized, where agile target acquisition and tracking are needed, and where real-time communications are needed during pointing, the telescope pointing and instrument pointing

options are preferred. If cryogenic temperature optics are needed, and gimbal mass/volume and cryocooler jitter are to be minimized, then the ability to transport either ambient or cryogenic

cooling across the gimbal is an absolute necessity. The solution to this problem requires two equally challenging developments: (1) a reliable ambient or cryogenic thermal transport system;

and (2) a reliable across-gimbal flex line system. Background information on these two topics is provided below. LHP Technology Options

Presented below is an overview of LHP technology followed by discussions of the operational principles, pumping ability, and design/operational constraints of a standard LHP as

well as a brief description of two other LHP-based alternatives. Overview. The loop heat pipe (LHP) is a revolutionary two-phase heat transport device needed to solve the gimbal heat transport problem. In the "standard" LHP configuration (two

other options are discussed below), it is a two-phase, liquid/vapor fluid loop with five basic components: evaporator, condenser, reservoir, transfer lines, and the working fluid. An LHP

reservoir is sometimes referred to as the "hydro-accumulator" or the "compensation chamber". Figure 2 illustrates a flow diagram of a standard LHP two-phase heat transport system. In a standard LHP, liquid is vaporized at the evaporator, which is mounted next to the heat source (the on-gimbal cryocoolers). The resulting vapor flows within small diameter tubing to the condenser, which is mounted next to the cooling source (the spacecraft-mounted radiator). The

vapor then condenses to liquid and flows back, again within small diameter tubing, to the evaporator. An additional wick structure spans the physical distance between the reservoir and the evaporator. This arrangement allows reliable loop start-up without preconditioning, it provides passive management of the working fluid (typically ammonia) over a range of operating

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temperatures, and it automatically regulates loop operating temperature over a wide range of

source/sink operating conditions. Operational Principles. Based on heat pipe principles, fluid circulation in an LHP is driven by surface tension forces (capillary action) that create very small pressure differences in the liquid and vapor phases. Like the heat pipe, the LHP is a completely passive device with no moving parts to wear out. When heat is added to the surface of a porous "wick" located in the evaporator, liquid is vaporized and capillary action acts to replenish the vaporized fluid. To accomplish this fluid replenishment, capillary action "pulls" liquid back into the evaporator. At the same time, at the vapor outlet of the evaporator, the just-vaporized vapor "pushes" the previously generated vapor towards the condenser. In this way, fluid is circulated in a loop and heat is transported from the source to the sink. One significant advantage the LHP has over the

conventional heat pipe stems from the small, isolated wick in the LHP evaporator, which enables the use of very small-diameter flexible tubing between the evaporator and condenser. In a conventional heat pipe, a continuous evaporator-to-condenser wick (typically narrow slots or grooves) necessitates the use of thicker-walled, not readily flexible tubing. This advantage also translates into increased pumping ability of the LHP because smaller pore-size wicks can be used.

Pumping Ability. The fluid pumping ability of a standard LHP (and the two other LHP options described below) increases as the pore size of the porous wick is reduced. At 300 K, an

ammonia-filled LHP with a wick can overcome an adverse elevation of over 6 meters of liquid ammonia. This pumping capability can also overcome accelerations due to gimbal starting/turning while maintaining significant heat transport capability. The ability to use very small pore size wicks in LHPs stems from the fact that LHP wicks are positioned at just one specific location in the evaporator, which minimizes viscous pressure drop in the wick structure. In all cases where pumping against gravity or external forces is required, the LHP is the preferred heat transport solution. The system payoff for having against-gravity pumping capability is that an LHP-enabled gimbaled spacecraft can most likely be tested in any orientation. Thus, the LHP is the only reasonable option for across-gimbal heat transfer. Design/Operational Constraints. In the standard LHP, the condenser must return subcooled liquid to the evaporator. Often, the subcooler is a separate heat exchanger on the space-facing radiator downstream of the condenser. Subcooling is required to compensate for

ambient heat absorbed by the liquid return line, ambient heat absorbed by the reservoir, and gimbaled-payload waste heat conducted across the wick structure. Sufficient subcooling must be provided to balance these heat inputs, otherwise the loop operating temperature will rise (i.e., the loop will "autoregulate") until an energy balance is achieved. Typically, it is desirable to

minimize these heat inputs in order to achieve a minimum operating temperatures. It is also necessary to separate the liquid return line from the vapor transport line across the gimbaled joint. Finally, it is necessary that the condenser (i.e., the radiator) be configured to provide optimum subcooling as required to achieve the desired system operating temperature. For the systems considered herein, these design constraints mean the following. Sufficient isolation of the liquid return line is needed considering the probable long line lengths, the fact that (relative to rigid lines) flex line volume is comparatively large, the fact that mechanical constraints required to achieve the needed flexibility across the joint may result in significant parasitic heat input, and the fact that there may be large heat leaks into a reservoir that is not "cold-biased." Thus, in a standard LHP, the reservoir must be either exposed to a cold environment (i.e., space) or protected from a hot environment. Thus, with a standard LHP, the reservoir and evaporator should not be wrapped together within a single MLI blanket.

Thermoelectric Option. As indicated, the standard LHP design approach requires two separate lines across the gimbal per LHP loop. One novel design alternative, as illustrated in Figure 3, is to mount a thermoelectric cooler (TEC) onto the reservoir to provide the cold-biased environment. The heat dissipated by the TEC is added to the normal evaporator heat load with a thermal strap. What this design option does is remove all constraints on LHP evaporator and


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reservoir placement. The reservoir no longer needs to be cold-biased or protected from a hot environment. In addition, the liquid and vapor lines no longer need to be thermally separated. A coaxial arrangement, where the liquid return line flows inside the larger vapor outer line, is the plumbing system of choice. Because of this coax arrangement, the liquid line can be a highly flexible (non-hermetically sealed) material like Teflon. This plumbing approach has been successfully demonstrated by Swales Aerospace on the HOST flight experiment. The objective there was to reduce the diameter and improve the flexibility of the transport line bundle in an LHP-like system that will ultimately be installed on the Hubble Space Telescope during Servicing Mission 4 (SM-4). Vapor Flush Option. Another novel design alternative for solving the problem addressed in this paper is the LHP vapor flush design option, which is commonly referred to as the advanced LHP or ALHP (see Hoang2). The ALHP flow diagram is provided in Figure 4. As indicated, there are actually two evaporators in an ALHP. The first evaporator, located at the heat source (the on-gimbal cryocooler), is a reservoir-less LHP evaporator. The second evaporator, located at the cooling source (the radiator), is a typical LHP evaporator and reservoir. The ALHP uses two superimposed flow loops. The first loop, denoted as the "steady-state" loop, carries the brunt of the cryocooler heat load. The second loop, denoted as the "vapor flush" loop, carries as much heat as is necessary to flush vapor from the on-gimbal evaporator. A small amount of heater power (typically about 5% of the total dissipation) must be added to the on-radiator evaporator, in order to provide the needed vapor flush mass flow. The ALHP has three important operational advantages over the standard LHP configuration, and two advantages over the TEC option. First, although there are three across-gimbal lines per loop, a tri-axial arrangement of the three lines is possible. Thus, the three lines can be coalesced into a single across-gimbal transport line, similar to the arrangement described above for the TEC option. Second, the LHP reservoir has been moved off-gimbal, thus the ALHP has less on-gimbal weight (perhaps 1 kg per LHP loop) than the TEC or standard LHP options. Third, the vapor flushing action of the on-radiator evaporator/reservoir system significantly reduces the need for subcooling. Thus, the radiator can be made more compact than it can in either other LHP option. The ALHP option was developed by TTH, Inc. The ALHP has been successfully demonstrated in laboratory testing.

Figure 2. Standard Loop Heat Pipe (LHP) Flow Diagram.

Figure 3. Loop Heat Pipe with Thermoelectric Cooler (TEC).

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Figure 4. Advanced Loop Heat Pipe (ALHP) Flow Diagram.

Flex Line Technology As described above, LHP technology solves a portion of the gimbal thermal transport problem. However, additional technology development was needed in two areas: (1) small diameter flexible lines; and (2) 2-axis flex line routing schemes. In the former technology area,

flex lines were needed with an internal diameter of 0.32 cm (0.125 inch) or less that could contain an internal pressure of at least 4.2 MPa (615 psi), which is ammonia's vapor pressure at 353 K, a worst-case hot temperature for STS payloads. The basic problem was that stainless steel bellows tubing is typically not available in diameters less than 0.64 cm (0.25 inch). In the latter technology area, bellows tubing must be flexed primarily in just one plane, so 2-axis flexibility can be best accomplished by separating the flex line into two separate flexible sections. To

address these two needs, a preliminary search for available flex lines and a preliminary study of possible line routing schemes were both carried out prior to the GATTS program. Small Diameter Flex Lines. During the preliminary flex line search, just one supplier of

reasonably small diameter stainless steel bellows tubing was found. This particular bellows tubing is a helical, seamless design with an ID of 0.39 cm (0.155 inch) and an OD of 0.66 cm (0.26 inch). Figure 5 illustrates an 18 cm long section of this tubing with brazed-on end fittings. Proof tests showed that that this tubing would not deform at pressures up to 10 MPa (1500 psi). Flex Line Routing. During the preliminary study to evaluate across-gimbal line routing schemes, the basic concept that was developed is shown in Figure 6. As indicated, the azimuth

axis flexible section consisted of 2 circumferential wraps (of a 3-line bundle) to give 360° of flexibility and the elevation axis flexible section consisted of 1 circumferential wrap to give 180° of flexibility. Rigid lines were used elsewhere. The 2-axis gimbal used to evaluate this line routing scheme was constructed from drawings provided by the AFRL. This gimbal was designated as the Gimbal Demonstration Unit (GDU). With this set-up, tests were conducted

examining the effects of line-induced torque and internal pressurization. During the GATTS program, this line routing scheme was refined to reduce line length and improve repeatability.

Figure 5. Flexible Bellows Tubing Sample for 2-Axis Gimbal Thermal Transport.


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Figure 6. Preliminary Line Management Scheme for 2-Axis Gimbal Thermal Transport.

Preliminary testing of line-induced torque with the configuration illustrated above indicated that each flexible line added about 0.14 m-N (20 in-oz) of torque in each axis, so a 3-line bundle added about 0.42 m-N (60 in-oz) of torque to each axis of the gimbal. Preliminary testing of system pressurization effects with the configuration illustrate above indicated that torque was independent of pressure to well above 0.7 MPa (100 psi), which would be the approximate pressure of an ammonia-charged LHP operating near room temperature. PROGRAM DESCRIPTION This section of the paper described the GATTS program in detail including the program objectives, requirements/specifications, design calculations/trades, design approach, test plans, and test results to date. Program Objectives The goals of the GATTS program are to analyze, design, manufacture, and test a full-scale, 2-axis gimbal with an accompanying flexible ambient LHP thermal transport system. The CRYOBUS 2-axis Gimbal Development Unit (GDU) is the starting point for the program. The plan is to incorporate motors, position encoders, torque measurement instrumentation, temperature sensors, and cryocooler simulators (heaters) onto the GDU. The goal is a motorized system whose axes can be independently slewed over their entire range of motion during testing. To manage line motion during flexing, the goal was to design a flight-like flex line routing and guide system, an appropriately sized flexible standard LHP heat transport system, and the required ground support equipment (GSE) for testing. The ultimate goal is to test the assembled gimbal LHP heat transport system continuously for 2-4 weeks to identify any unknown dynamic or potentially deleterious effect on the system's ability to meet the mission requirements outlined below. The system will be tested in air under ambient laboratory conditions.

Requirements/Specifications The basic requirements/specifications for the GATTS program are listed in Table 1. These were prescribed and/or derived by AFRL and Swales as representative of a range of missions requiring across-gimbal ambient thermal transport capabilities. Design Calculations/Trades A range of design calculations and trades were done to determine the test orientation limitations (if any), bellows line diameter and length, bellows line pressure handling capability, optimal evaporator placement, and motor torque requirements.

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Test Orientation Limitations. One of the primary advantages of the LHP over conventional heat pipes is the strong capillary pumping action in the evaporator. Because

capillary pressure is inversely proportional to the effective wick pore radius, this pumping ability increases as the wick pore size is reduced. In fact, in many cases virtually all ground test orientation limitations can be eliminated. To achieve this favorable result, the maximum against-

gravity pumping height (H) must exceed the physical separation distance between the condenser and evaporator. With currently achievable evaporator pore radii of about

with nickel wicks,

LHP evaporators with ammonia as the working fluid can overcome 6 meters of adverse elevation. Equations (1-2) illustrate the relationships between surface tension pore radius liquid density and H. Thus, GATTS and most small-to-medium sized systems using an across-gimbal LHP can probably be tested in any orientation. One final factor to consider is pressure drop in the lines, and this topic is discussed next.

Line Diameter and Length. To reduce line length and diameter, maximize line flexibility, minimize torque, and minimize pressure drop, calculations were carried out for the flexible tubing to determine pressure drop as a function of line diameter and length. The results of the calculations are as follows. With ammonia as the working fluid, the pressure drop for a 200 W heat load within a 1.25 m long section of the selected flex line was about 5% of the available capillary pumping head and about 7 times higher than that of smooth-walled tubing. Bellows Line Pressure Containment With ammonia as the working fluid, system pressures will range from 0.54 MPa (80 psi) at 280 K to about 1.4 MPa (200 psi) at 310 K. This working pressure range will pose no difficulty for the selected flex line given its proof test results

presented earlier. Evaporator Placement Considerations. During gimbal movement, the fluid within the LHP will be subject to accelerations. To minimize the impact of these accelerations on LHP


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operation, the evaporator will be located as close as possible to the axis of rotation of each axis. Test results are available which indicate that accelerations of up to 5g will not cause deleterious

effects in LHP systems. As a further precaution, the centerlines of the LHP reservoir and evaporator (typically coincident with each other) will be parallel with the elevation axis of rotation. When oriented nominally (see Figure 6) or upside down, this placement will prevent the reservoir from being below the evaporator with respect to gravity (although certain sideways orientations may be problematic). Additionally, the reservoir will be radially inboard of the evaporator with respect to the azimuth axis, which will force fluid in the reservoir to feed into the evaporator as the azimuth axis rotates in either direction. Motor Torque/Placement and Axis Motion. To ensure the procured motors are located optimally and have sufficient torque to drive each GDU axis enough to achieve the maximum slewing rates listed in Table 1, a design study to quantify system torque and evaluate optimal motor placement was implemented. The results of this design study are provided below. The elevation motor will be mounted to the elevation on the opposite side of the gimbal from the flex lines. It will be capable of producing approximate 0.56 m-N of torque, which includes about 0.22 m-N of margin. The planned motion for the elevation axis is to accelerate linearly at from –90° to 0° and then decelerate at from 0° to +90°. This acceleration-deceleration profile will yield a maximum angular velocity of about 200 deg/s. The time required for a complete –90° to +90° elevation slew is about 1.8 seconds. The azimuth motor will be mounted to the azimuth axis underneath the azimuth bearing. It will be capable of producing approximately 2.2 m-N of torque, which includes a margin of about 0.56 m-N The planned motion for the azimuth axis is to accelerate linearly at from –200° to 0° and then decelerate at from 0° to +200°. This motion will yield a maximum angular velocity of around 280 deg/s. It may be preferable to modify this plan so that the motor maintains a constant velocity once an angular velocity of 200

deg/s is achieved. The time required for a complete –200° to +200° slew is about 2.8 seconds. Design Approach

The primary goals in designing the GATTS test system included but were not limited to the following: (a) maximize the repeatability/reliability of the flex line system; (b) minimize the flex line and total line lengths; (c) minimize the line-induced torque, (d) minimize the weight of the on-gimbal LHP, and (e) ensure sufficient instrumentation/test monitoring equipment to answer the major technology question of whether gimbal motion affects LHP operation. An equally important technology question, which will be part of the GATTS follow-on program assuming this effort is successful, is whether LHP flow in conjunction with gimbal motion introduces excessive jitter into the optical/sensor system. In the interest of brevity, the remainder of this section just addresses the issues of line management and length minimization. Line Management and Length Minimization. From prior testing, the line routing scheme developed prior to GATTS (shown in Figure 6) had excessive line length, questionable launch stability, and uncontrolled line movement in both axes. To rectify these deficiencies, a study was carried out to look at alternative line routing methods. In this study, the schemes all centered on the concept of "minimum dynamic bend radius". The minimum dynamic bend radius specified by the manufacturer of the selected flex line is 7.5 cm (3 inches). The line management concepts that evolved from the study are illustrated schematically in Figure 7. In the elevation axis routing method, only 20 cm (8 inches) of flex line are used. One of the nice features of this routing method is that the flex line is not flexed (or stressed) in its neutral position at 0°. This feature means that there will be very good line controllability during both launch and on-orbit operation. In the azimuth axis, because of the much greater range of motion required (i.e., more than one complete revolution is needed), the flex line needs to be about 100 cm (40 inches) in length. One unique feature of the azimuth line routing method, which avoids having the lines rub against one another, is the helical wrap shown in Figure 7. Although this method means that the lines will not bend strictly in plane, the slight out-of-plane

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bending does not appear to be a problem. However, compared to the elevation axis line routing method, the azimuth axis will probably require a more substantial launch lock subsystem to prevent the lines from moving excessively during launch as well as a low-torque line management restraint subsystem to ensure repeatable wrapping and unwrapping of the lines during operation. The launch restraint subsystems will be developed as part of the GATTS follow-on program, and will not be required for this effort. Test Plans

This section of the paper addresses the plans for the test set-up, provides a preliminary test matrix, and discusses the anticipated results. Again, at the time this paper was written, the test hardware was still in the design stage. Test Set-up. The test set-up for the GATTS program will be similar to the configuration illustrated in Figure 8. Plans for modifying the existing GDU to meet GATTS program goals

include the following: (1) the GDU will be modified with the addition of motors on the azimuth and elevation axes; (2) a single flexible LHP system will be utilized (although the GATTS system will be designed to accommodate an additional flexible LHP system should follow-on work for a redundant flight system be warranted); and (3) the LHP evaporator and reservoir will be aligned along the elevation axis with the reservoir inboard of the evaporator. This evaporatorto-reservoir orientation ensures that azimuth gimbal movement always forces reservoir fluid into the evaporator.

Test Matrix. At the time of this writing, three primary sets of tests are planned for the

gimbal thermal transport system. The first will be an accelerated limited life test with the system pressurized with nitrogen to identify any unanticipated lifetime issues with respect to the line routing scheme that might cause the system to fail prematurely. Once this test has demonstrated sufficient lifetime in each axis (e.g., greater than 50,000 cycles or 2.5 times the requirement of 20,000 cycles listed in Table 1), a new set of flex lines will be added and the system will be charged with ammonia for the second set of tests. In this second set of tests, LHP system performance will be generated without gimbal motion for use as a comparison baseline. Finally, in the third set of tests, the GATTS LHP will be tested with simultaneous gimbal motion in each axis. The types of subtests that will comprise this third set of tests has not been determined. Ideally, all sorts of combinations of axis motions will be investigated to identify system limitations (if any). Anticipated Results. Based on the favorable results presented by

who successfully

tested an LHP under accelerations up to 4.7g, it is anticipated that the LHP system will be able to

Figure 7. Illustration of GATTS Flex Line Management/Length Minimization Approach.


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Figure 8. Test Set-Up for the GATTS Program.

operate successfully during gimbal motion. The maximum accelerations in the GATTS system will be about 0.5 g (in each axis), hence gimbal motion should not affect LHP operation. CONCLUSIONS This paper has presented the rationale, technical background, and current plans for the development and testing of an across-gimbal ambient thermal transport system. The application for this system is for cryocooled infrared payloads for space-based DoD and NASA missions that require advanced imaging and tracking capabilities. The status of this development effort, known as GATTS (for gimbal ambient thermal transport system), is that the test system is still in the design phase. However, no problems in successfully achieving the program goals are foreseen. In about 2 months from the time this paper was written (in June 2000), manufacture and assembly of the test hardware will have been completed. Testing will commence shortly thereafter and test results should be available another month or so after that. When available, test results will be published pending approval by the Air Force. ACKNOWLEDGMENT The authors would like to gratefully acknowledge the funding, program management, and technical management support provided by the Air Force Research Laboratory for the work performed in this effort, which was part of a broader research program known as CRYOBUS. REFERENCES 1. Bugby, D., Brennan, P., et al., "Development of an Integrated Cryogenic Bus for Spacecraft Applications," Space Technology and Applications International Forum (STAIF-97), January 1997. 2. Hoang, T., Kim, J., and Cheung, K., "Design and Test of a Proof-of-Concept Advanced Capillary

Pumped Loop," SAE 27th International Conference on Environmental Systems, July 1997. 3. Ku, J., Kaya, T., et al., "Testing of a Loop Heat Pipe Subjected to Variable Accelerating Forces," Spacecraft Thermal Control Technology Workshop, The Aerospace Corporation, El Segundo, CA, March, 2000.

Cryocooler Interface System G.S. Willen

Technology Applications, Inc. Boulder, CO, USA 80301 B.J. Tomlinson Air Force Research Laboratory

Kirtland AFB, MM, USA 87117 ABSTRACT

For actively cooled cryogenic systems it is usually necessary to locate the cryocooler in close proximity to the cooled assembly. This places highly demanding requirements on the cryocooler, its integration into the spacecraft, and introduces unwanted electrical, magnetic, and mechanical

disturbances. A unique Cryocooler Interface System is being developed in which the cryocooler can be remotely located from the cooled elements, virtually eliminating cryocooler disturbances.

This system accommodates cooling across gimballed axes, provides nearly constant temperature cooling under variable loads, can cool large area and/or distributed elements, and simplifies cooling system integration. The capability to provide cooling across a gimbal axis is important in advanced surveillance systems such as the Spaced Based IR System-Low (SBIRS-Low) Segment, a near-term Air Force program for developing and deploying a constellation of low-earth orbiting observation satellites that incorporate cooled optics mounted on a two-axis gimbal. Mounting the cryocooler on the gimbal has a number of drawbacks; the most serious are weight, vibration, and heat rejection. To address these issues, a Cryocooler Interface System is being developed under an AFRL/VSSS Phase II Small Business Innovative Research (SBIR) program (Contract No. F29601-99-0009) for cooling the gimbal-mounted optics on SBIRS-Low. INTRODUCTION

The emphasis on aerospace cryocooler development has focused on reliability, efficiency, and vibration reduction. These goals have been generally realized; however, many practical issues concerning cryocooler integration remain unresolved. Some of the more critical issues are: 1) providing cooling across gimbal axes, 2) large area or distributed cooling, 3) effective cryocooler heat rejection, and 4) overall cooling system integration. Technology Applications, Inc. (TAI) is developing a unique Cryocooler Interface System (CIS) in which the cryocooler can be located remotely from the cooled elements. The CIS accommodates cooling across gimbaled axes, provides near constant temperature cooling under variable loads, and greatly simplifies cooling system integration. It also allows cooling of large area and/or distributed elements, virtually eliminates cryocooler disturbances, is insensitive to gravity level, offers ease of redundancy, and can be used with most types of cryocoolers. Mounting a cryocooler on-gimbal adds weight, increases inertia, is a source of vibration, and requires routing power and control wires across the gimbal axes. Since cryocooler heat absorption Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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is highly localized, high conductivity paths need to be incorporated into the gimbal mounted optical system, adding weight and introducing temperature gradients. Perhaps the most critical issue is effectively rejecting cryocooler-generated heat. For heat loads of 10 to 15 watts, the amount of heat rejected is on the order of 100 to 150 watts; this requires a large radiator and, because the gimbal can be pointed in arbitrary directions, the radiator performance can vary as a function of its view to space. Conventional approaches for cooling large area focal plane or optical assemblies using thermal straps can induce large temperature gradients across the structure, are thermally inefficient, add weight, and can put excessive stress on the cryocooler cold tip. Since spacecraft heat rejection radiators are not typically located near the cooled elements, a solid conductor and/or heat pipes must be employed to transfer cryocooler-generated heat to the radiator. This not only adds weight, it increases the cryocooler internal temperature, thereby increasing input power and reducing cryocooler efficiency, reliability, and lifetime. Currently, the spacecraft system engineer must commit to a cooling system early in the design phase. As the design evolves and system requirements change, the impact upon the cooler and its integration into the spacecraft can lead to excessive cost and schedule growth, often compromising overall system performance. The CIS will provide the spacecraft system engineers with a high degree of cooling system design flexibility allowing them to design their systems without locking into a specific cryocooler and a location. SYSTEM DESCRIPTION

The CIS, shown in Figure 1, consists of a low-pressure fluid-cooling loop that absorbs heat from the optics and rejects it at the cryocooler; the major CIS components are identified by the bold call-outs. The fluid loop incorporates flexible segments that provide the range of motion for routing the coolant lines across the gimbal axis. All CIS components, except for the transfer line, can be located wherever it is most convenient; for the cryocooler, this will generally be in the proximity of the spacecraft heat rejection radiator. Once the optical bench heat exchanger, cryocooler, and circulation compressor assemblies are integrated into the spacecraft, the transfer lines are installed connecting the two assemblies. Any changes in the spacecraft design will primarily affect only the transfer line routing. The key CIS design requirements are given in

Table 1.

Figure 1. Cryocooler Interface System concept provides system flexibility.

CRYOCOOLER INTERFACE SYSTEM

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Full cooling system redundancy is easily accommodated without the need for thermal switches, as the coolant loop for the non-operating cryocooler adds only a few tens of milliwatts parasitic heat load. If full redundancy is not required, several levels of redundancy can be

implemented wherein only the most failure prone components such as the cryocooler, compressors, and electronic control are duplicated.

The Interface System operation is illustrated by following the path of the working fluid through the block diagram in Figure 2, and corresponding state points on the TemperatureEntropy diagram in Figure 3. For an optics temperature of 110 K, the circulating coolant is methane, which has a normal boiling point of 111.6 K and is ideal for cooling in the 100 K to 120 K range. The methane return gas is compressed adiabatically (1-2) to a pressure slightly higher than the saturation pressure corresponding to the cryocooler cold tip temperature (for 120 K, the methane pressure must be psia). The heat of compression is removed by cooling the methane to near ambient temperature with a simple heat exchanger attached to the spacecraft heat rejection bus (2-3). Next, the cold return gas from the optics precools the high-pressure methane stream in a counter-flow heat exchanger (3-4). The methane is then condensed (4-5) and subcooled (5-6) at constant pressure as it passes through the cold tip heat exchanger. The subcooled liquid methane flows through a rigid, insulated transfer line that is connected to the optics via flexible, low-stiffness line sections designed to route the coolant across the gimbal axes (6-7). It is then expanded through a throttle valve to a pressure of 12 psia and a corresponding temperature of 109 K (7-8) and routed through a heat exchanger thermally attached to the optical bench (8-9) where the liquid phase boils at constant temperature, absorbing the heat load. After the methane exits the optical bench heat exchanger, it is routed through the return transfer line (9-10), through the counter-flow heat exchanger (10-1), and into the compressor inlet. Since the optical bench heat exchanger temperature is determined by the coolant loop pressure, the cryocooler operating temperature can be 5-to-10 degrees above the optics temperature and vary as much as This reduces both the input power and temperature control requirements imposed upon cryocoolers used in conventional cooling system approaches. Gas storage volumes are located on both the high- and low-pressure sides of the compressor. Each is sized to maintain the absolute pressure in the system within safe limits when the system is warmed to ambient temperature, and to buffer against optics heat exchanger temperature variations that result from system pressure variations due to variable heat loads.

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Figure 2. Cryocooler Interface System functional block diagram.

Figure 3. Cryocooler Interface System thermodynamic cycle.

It is not desirable to periodically vary the cryocooler heat lift to match the large SBIRS-Low load variations, nor is it efficient to always operate the cryocooler at maximum load, throwing away the excess cooling. A unique feature of the CIS is its ability to be easily integrated with an external phase-change thermal storage unit (TSU). For SBIRS-Low, a TSU using 2methylpentane as the phase change storage medium is a promising option. Since 2methylpentane freezes at 119.6 K, yet remains liquid up to 60°C, there is no need for a large, external expansion chamber. However, because 2-methylpentane freezes at 119.6 K, the

methane supply pressure must be psia to subcool the liquid methane. The circulation system compressor is the only active component in the system. It has a volumetric flow rate of three standard liters per minute and a pressure ratio of about 4:1. TAI is developing a long-lifetime rotary compressor for the CIS program. The cryocooler cold tip heat exchanger is a small diameter stainless steel tube wrapped around a copper spool that is attached to the cryocooler cold tip. If a TSU is used, the cold-tip heat exchanger will be integrated into the TSU. Since the coolant flow through cold tip heat exchanger is mostly liquid, its effectiveness is greater than 0.98. The throttle valve is located downstream of the cryocooler and reduces the liquid methane pressure to approximately 12 psia, corresponding to a saturation temperature of 109 K. A capillary tube was chosen because, for the same pressure drop, its diameter is considerably larger than an equivalent orifice, making it less sensitive to contamination. It is also more stable, as the flowrate changes only slightly with temperature, whereas, orifice flow is very sensitive to temperature changes.


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The optical bench heat exchanger consists of a small diameter tube attached to the optical bench; it is designed to minimize spatial and temporal temperature gradients. In a zero-gravity environment, the two-phase methane will continuously wet the tube walls. Bending the coolant

tubes in a serpentine shape will assure the liquid in the flow stream continuously contacts the tube walls. Filter assemblies are incorporated into the circulation system to minimize the potential for throttle valve plugging. Getters cannot be used in the CIS since the methane coolant is strongly adsorbed by the getters. Of the potential outgassing contaminates, only water and carbon dioxide will condense; their freezing temperatures are 273.3 K and 216.5 K, respectively. Prior to charging the system with methane, the CIS will be thoroughly cleaned, purged with hot nitrogen, and evacuated. What little water and carbon dioxide are left will, for the most part, freeze out on the counterflow and cold-tip heat exchangers. In the event the throttle valve becomes plugged, a defrost heater installed on the throttle valve can be activated to thaw any potential plugging resulting from condensibles. The methane transfer line is composed of flexible sections that cross the gimbal axis connected by rigid sections. One of the most demanding requirements is transferring coolant across the gimbal axis. To accommodate the gimbal motions, the gimbal axis transfer line assembly incorporates low-stiffness, flexible sections that cross the azimuth and elevation axes. These flexible transfer lines are small diameter 300 series stainless steel tubes formed into a helical configuration, winding and unwinding as the axes rotate. The azimuth axis rotates between the elevation axis rotates from 0° to 90° and back. The primary issues are achieving the full range of motions, minimizing torque on the gimbal, and meeting the 400,000-cycle lifetime requirement. Since the transfer lines constitute the major

heat leak into the system, and every milliwatt of heat leak translates into a milliwatt of additional load on the cryocooler; thus, the lines are gold coated to minimize heat leak. The gimbal axis transfer line assembly is shown in Figure 4.

Figure 4. Gimbal axis line routing.

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COMPONENT DEVELOPMENT TESTING Component development testing of the azimuth and elevation axis flexible line segments and the breadboard methane circulation system has been ongoing since Phase I of the SBIR program. The circulation compressor is scheduled for performance and life testing starting in July 2000. Full-scale models of the azimuth and elevation axis gimbal joints were constructed using programmable stepper motors to provide the full range of motion. Representative flexible line segments were fabricated and tested; the test set-up is shown in Figure 5. Preliminary azimuth axis testing, conducted during the Phase I program, demonstrated over cycles of The elevation axis flexible segment, consisting of both a supply and return line, was cycled over times (0° to 90° and back) with no signs of fatigue or wear. One complete gimbal axis cycle every 100-minute orbit equates to 5,260 cycles per year; the 2.5 million cycles demonstrated is equivalent to 475 years. The elevation axis torque was measured prior to and after completion of testing; the torque was oz-in for both cases. This torque value is an average of six measurements; it represents the net torque after the bearing torque was subtracted. All tests were conducted at room temperature. Since the modulus of elasticity of 300 series stainless steel increases about 20% at 110 K, the gimbal torque will increase slightly at operating temperatures. The azimuth axis coil has a larger diameter than the elevation axis coil; therefore, its net torque is expected to be less than 3 oz-in. The methane circulation system was tested to demonstrate the CIS remote cooling capability

and assess its operating characteristics. This test set-up and the accompanying circulation system

schematic are shown in Figure 6. The circulation system dewar assembly includes a liquid reservoir that is used in place of a cryocooler to absorb the circulation system heat load, the counterflow and cold-tip heat exchangers, and the capillary throttle valve (refer to the components enclosed by the dashed box). The transfer line from the throttle valve to the load end is approximately one-meter long. The load-end heat exchanger is a one-meter long tube coiled around a one-inch diameter copper mandrel with an imbedded cartridge heater that supplies the variable cold-end heat load. A heater is attached to the reservoir cold tip, and a temperature controller is used to maintain the methane temperature exiting the cold-tip heat exchanger at

nitrogen

Figure 5. Azimuth and elevation flexible line segment test set-ups.


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Figure 6. Methane circulation system test set-up.

Typical circulation system performance is illustrated by the temperature profiles in Figure 7. The upper temperature data are the cold-tip outlet and throttle valve inlet temperatures (T7 and

T2). The lower temperature data are the throttle valve outlet, and the load heat exchanger inlet and outlet temperatures (T3, T4, and T5). At the start of testing, the circulation system was stabilized at a nominal 8 watts cold-end heat load. The cold-end heater power was then varied between six and twelve watts. As shown by Figure 7, the load-end temperature generally stayed within over the load variation, except at the high and low load extremes. The conductance between the reservoir and cold-tip heat exchanger was sufficiently low that, for a cold-end load of 12 watts, the methane temperature exiting the cold-tip heat exchanger started rising above 119 K and was no longer subcooled. In response, the load-end temperature increased 3 K above the 109 K nominal target temperature. When the cold-end load was lowered to 6 watts, the cold-tip heater did not have sufficient power to maintain the methane temperature at 118K, consequently, the methane was highly subcooled and the load-end temperature dropped to 106 K.

The cold-end temperature variation with load observed in this test can be attributed primarily to the test set-up, which was designed for a maximum load variation between 6 and 12 watts. A commercial linear compressor was used; this limited the supply pressure to about 28 psia at a flow rate of 30 mg/s. In addition, the cold-end heat exchanger has very little thermal mass. An operational system will use a circulation compressor sized for the load and the optics subsystem, having a much higher thermal mass, will greatly attenuate any temperature variation.

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Figure 7. Circulation system test results.

PROTOTYPE INTERFACE SYSTEM TESTING Testing of the prototype CIS under simulated operational environments will be conducted using the test set-up illustrated in Figure 8. The test set-up will be instrumented to provide the data necessary to fully characterize the CIS performance and operation; the instrumentation and their locations are defined by the system schematic in Figure 9. A representative set of gimbal axis motions and optics heat loads will be used in the prototype test to simulate a realistic mission scenario. By cycling the gimbal once a minute, a ten-year onorbit simulation can be completed in approximately five weeks. The test will be performed in a

vacuum

torr, and at room temperature.

Figure 8. Prototype Interface System test set-up.


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SUMMARY AND CONCLUSIONS A unique Cryocooler Interface System is being developed in which the cryocooler can be remotely located from the cooled elements. Its capability to provide cooling across a gimbal axis is important in advanced surveillance systems such as cooling the SBIRS-Low optics. Critical component tests have demonstrated both the capability for effectively transferring coolant across the gimbal axes and the operational stability of the circulation system. Prototype system testing will demonstrate the overall system operation and capabilities. ACKNOWLEDGMENTS

The authors wish to acknowledge the support of TAI personnel Steve Nieczkoski and Edward Myers for their considerable contributions in the design, analysis, and testing of the Cryocooler

Interface System.

Development and Testing of a High Performance Cryogenic Thermal Switch B. Marland, D. Bugby, and C. Stouffer

Swales Aerospace Beltsville, MD, USA 20705

B. Tomlinson and T. Davis Air Force Research Laboratory Kirtland AFB, NM, USA 87119

ABSTRACT

This paper presents development details and performance test results of a high performance cryogenic thermal switch (CTSW) for coupling redundant cryocoolers to cryogenic components

with minimal off-cooler parasitics. Because gas-gap, hydride-pumped CTSW designs have not reliably met performance goals of an "on" resistance less than 2 K/W and an "off" resistance greater than 1000 K/W, a simpler, more reliable device was sought. The device that was developed is based on the reversible and highly reliable physical process of differential thermal contraction/expansion of stainless steel relative to beryllium. The 250 gram Swales differential coefficient of thermal expansion (CTE) CTSW (or SDCC) has just 3 machined parts: two cylindrical beryllium discs (same diameter, different lengths) and a thin-walled stainless steel tube. In ground testing, the SDCC demonstrated an "off" resistance of 1400 K/W and an "on" resistance of 1.2-2.0 K/W over a cold end temperature range of 25-50 K and a warm end temperature range of 230-300 K. Other issues addressed in the paper include alternative gas-gap (non hydride-pumped) designs, the option of "cross-strapping" and CTSW reliability. Finally, an advanced SDCC design is also briefly described that can reduce CTSW mass to just 50 grams with virtually no (surface area-induced) parasitic heat input into the cryogenic system. INTRODUCTION The reliability of today's cryocoolers remains a limiting design constraint for many longlife, low risk cryogenic space applications. As a result, near term cryocooler space applications will probably require cryocooler redundancy, which is accompanied by an additional parasitic heat load from the non-operating cryocooler. For typical space cryocoolers operating without a thermal switching device, the parasitic load stems primarily from conduction through the nonoperating cyrocooler expander. The thermal resistance of this conductive path is generally 400500 K/W for space cryocoolers. As a result, the parasitic load due to the non-operating cryocooler is approximately 0.5 W ([285-60]/450) at 60 K, Bugby1.

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In order to minimize cooling and input power requirements, a reliable cryogenic thermal switch is desirable. A properly designed CTSW increases the thermal isolation between the instrument and expander body of the non-operating cryocooler and reduces the parasitic load from the non-operating cryocooler by 65-80%, Bugby1. As a result, cryocooler cooling and input power requirements are substantially reduced. The benefit of a CTSW becomes increasingly pronounced as the instrument operating temperature and cryocooler efficiency are decreased. From 10-30 K, the cooling and input power requirements needed to overcome an additional parasitic heat load of 0.5 W are prohibitive. In this operating regime, a CTSW becomes essential. Figure 1 illustrates a redundantly cooled dual CTSW system. By reducing the parasitic penalty for each non-operating cooler, CTSWs invite the thermal systems engineer to consider the use of multiple redundant coolers, lower reliability coolers and low-cost tactical coolers. Figure 2 illustrates the use of multiple CTSWs in conjunction with an array of low-cost tactical coolers.

CRYOGENIC THERMAL SWITCH PERFORMANCE REQUIREMENTS

The development of the cryogenic thermal switches presented in this paper was funded by an AFRL-sponsored initiative to incorporate new and enabling cryogenic technologies into space

systems. This initiative, dubbed the Integrated Cryogenic Bus (ICB), endeavors to combine a

Figure 1. Dual CTSW System.

Figure 2. CTSW/Low-Cost Cryocooler System.

DEVELOPMENT OF A CRYOGENIC THERMAL SWITCH

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range of cryogenic integration solutions to meet the needs of future space applications. The design of the SDCC is driven by the ICB established CTSW performance requirements, outlined in Table 1. SWALES DIFFERENTIAL THERMAL EXPANSION CTSW (SDCC) OVERVIEW

In an effort to meet the ICB performance requirements while providing a high reliability CTSW, a novel and promising concept was invented. The SDCC design, which has just three parts, is based on the reversible and highly reliable physical process of differential thermal contraction/expansion of stainless steel relative to beryllium. The SDCC design is an evolution of a gas-gap design developed by Swales and flown on STS-95 as part of the CRYOTSU

Hitchhiker Flight Experiment in October 1998. The Swales gas-gap design demonstrated the ICB performance requirements in ground testing but due to the impressive potential of the SDCC and system level reliability concerns with the gas-gap CTSW (which will be elaborated upon in subsequent sections) emphasis for CTSW development shifted to the SDCC design. Unlike gasgap CTSW designs, the SDCC design does not utilize a working fluid. As a result, no hermetic seals or getter/hydride pump to actuate the switch are needed and switch reliability is greatly enhanced. Figure 3 illustrates the SDCC design. Swales Aerospace is in the process of patenting this design. The SDCC is actuated "on"/"off" by contracting/expanding the stainless steel tube relative to the beryllium discs. When the smaller beryllium disc is cooled sufficiently, the higher CTE of the stainless steel tube compared with the beryllium causes the narrow flat gap in the switch to close, turning the switch "on". The time needed to actuate the SDCC "on" is minimized by the low mass/thermal impedance of the smaller beryllium disc. For low CTE/beryllium instruments, the beryllium construction of the SDCC allows for direct mounting, thereby, reducing the number of flexible thermal links needed and greatly enhancing the effective "on" performance of the system.

By warming the smaller disc or a portion of the stainless steel tube, a temperature gradient in the tube is created, causing the tube to expand, the narrow gap to open and the SDCC to turn "off". For missions utilizing redundant or multiple cryocoolers, once a cryocooler fails or is turned off, it is desirable to thermally isolate that cooler from the cryogenic instrument as rapidly as possible. By temporarily applying power to a small heater (approximately 0.3 W was needed

for the prototype SDCC) on the center of the stainless steel tube, the tube expands and the SDCC is rapidly turned "off" ("on" to "off" actuation times are less than 5 minutes for the prototype unit). Once in the "off" condition, the small beryllium disc becomes thermal isolated and its temperature may be increased with heater power or allowed to increase due to parasitic heating.

Figure 3. SDCC Design.

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After the smaller disc warms sufficiently (approximately 230 K with the prototype SDCC), the SDCC will remain in the "off" condition without heater power and the smaller disc will operate at a temperature above 230 K (275 K during SDCC prototype testing). The SDCC has been operated at temperatures as low as 25 K but operation could be easily extended to below 4 K. SWALES DIFFERENTIAL THERMAL EXPANSION CTSW (SDCC) DESIGN

In addition to its novel three-piece construction and unique beryllium interface compatibility, the SDCC has several distinguishing design features, (1) the extremely narrow (0.07 mm nominally), highly precise flat gap that separates the beryllium cylindrical parts, (2) the thin-walled stainless steel tube which aligns the beryllium parts and provides structural support, (3) the large size difference between the two beryllium components and (4) the proprietary gold plating process applied to each beryllium part to maximize the contact conductance. As mentioned previously, a narrow flat gap separates the two beryllium parts. A major advantage of the flat gap SDCC design over other mechanical and gas-gap CTSW designs is that the gap width can be verified at any time, not just prior to final assembly. As a result, the flat gap design significantly reduces the risk of a thermal short when the switch is "off". In addition, by varying the gap width, the temperature at which the switch turns "on"/"off" may be adjusted. In order to maintain the narrow flat gap and provide the thermal switching mechanism of differential thermal contraction, a thin walled stainless steel tube precisely aligns the two beryllium cylindrical parts. Aside from providing critical alignment, the stainless steel tube represents the primary support structure in the SDCC. For a given cross sectional area, the thinwalled design (0.13 mm as shown in Figure 1) maximizes the bending stiffness of the tube and provides significantly more structural support than a solid rod with equal cross section. By mounting the larger beryllium part directly to the cryogenic component, the stainless steel tube only supports the mass of the smaller beryllium piece (which is minimized for structural and thermal response considerations) and a portion of the flexible cryocooler interface. The tube diameter is sized to provide enough bending stiffness to avoid mechanical contact of the gap faces when the smaller beryllium part is subjected to a 13 g lateral load, approximately 3 N. A robust fundamental lateral frequency of more than 170 Hz and positive stress margins with a 40 g lateral load result from this sizing. The structural design of the SDCC is more robust than its predecessor, the hydrogen gas-gap CTSW, which was successfully qualified and flown on STS-95 on the CRYOTSU Hitchhiker Flight Experiment. While the switch is "off", the thin walled construction of the stainless steel tube represents the primary thermal path from a non-operating cryocooler to the instrument, minimizing its cross sectional area is critical. The tube's conductive resistance of approximately 2000 K/W for the operating range described in Table 1 ensures effective thermal isolation while the switch is "off". During "on" operation, the relative differential contraction of the stainless steel tube relative to beryllium closes the gap and provides approximately 450 N (analytical estimate) of contact force between the gap surfaces. The flat gap design of the SDCC loads the stainless steel tube axially (its stiffest direction) ensuring large stress margins during "on" operation. Operationally, the large size difference between the two beryllium components also plays an important role. During "off" operation, the large temperature gradient along the length of the tube and the thermal contraction of the larger beryllium part maintain a gap in the system. The length of the larger beryllium component relative to the assembly length ensures sufficient beryllium thermal contraction to maintain the SDCC in the "off" condition without heater power on the tube. Finally, a proprietary gold plating process applied to the gap surfaces of each beryllium component reduces the contact resistance between these surfaces during "on" operation. The SDCC has been cycled over 30 times with the gap operating force of 450 N and an additional 20 times in thermal cycling without any evidence of surface degradation or cold-welding. Life-cycle testing awaits to fully qualify the SDCC for flight applications.


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SDCC TEST METHODOLOGY AND RESULTS

In order to accurately determine "on" and "off" performance of CTSWs, simple yet effective test methods were developed during the Swales gas-gap CTSW development effort for the STS95 Hitchhiker Flight Experiment. These methods, as they relate to the SDCC, will now be discussed. In order to conduct an "on" test with the SDCC, the switch is turned on by mounting the larger beryllium part to the cold head of a cryocooler and reducing the temperature of the surrounding environment until the switch turns "on". Once the switch is "on", the surrounding environment may be warmed and the switch will remain "on". When steady state temperatures in the switch and environment are reached, an initial temperature difference across the switch is measured where is the temperature of the warm end and is the temperature of the cold end). This step provides the datum from which the temperature (relative) measurement errors and the effects of external parasitics can be eliminated. Then, a known heater power is applied to the warm end and a new is measured. The "on" conductance is the heater power divided by the change in temperature difference. Eq. (1) provides the analytical relationship. The "off" test is conducted by first warming the surrounding environment to ambient conditions and then turning the switch "off" by applying power to the heater on the center of the tube. Once the smaller beryllium part reaches a critical temperature (approximately 230 K with the prototype SDCC), the tube heater may be turned off and the switch will remain "off". With the SDCC in the "off" condition, the "off" performance is easily determined. The required heater power needed to equilibrate the warm end and surrounding environment temperatures, is then applied to the warm end. This heater power "zeroes out" the parasitics. The "off" resistance is the resulting divided by the heater power. Eq. (2) illustrates the analytical relationship.

Figures 4, 5 and 6 demonstrate ground testing results for the SDCC. The test results are highlighted by an "on" resistance of 1.2 K/W at 50 K, 2.0 K/W at 25 K and an "off" resistance of

Figure 4. SDCC "ON"/"OFF" Performance.

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Figure 5. SDCC "ON" Performance 25-35 K.

Figure 6. SDCC "ON" Performance 45-55 K.

1400 K/W. The reduced "on" performance of the switch at decreased temperatures is attributed to the decreased thermal conductivity of the beryllium through the constricted heat path at the

contact interface. The rapid transitions from "off" to "on" (at 230 K) and from "on" to "off" as well as the stable behavior in both the "off" and "on" conditions should also be noted. In addition, it is evident from the data and the design descriptions above that the SDCC operating range may be easily extended to below 4 K.


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GAS-GAP CRYOGENIC THERMAL SWITCH OVERVIEW As previously discussed, Swales began CTSW development work under the ICB initiative.

A hydrogen gas-gap CTSW was developed and flown as one of four advanced cryogenic integration devices flown on the STS-95 CRYOTSU Hitchhiker flight experiment in October 1998 under this initiative. Figure 7 illustrates the layout of the CRYOTSU flight experiment. No

degradation in switch performance due to launch loads was observed upon comparison of preflight and post-flight laboratory test results. The concept for the hydrogen gas-gap cryogenic thermal switch utilizing a getter/hydride

pump is based on a JPL design, Johnson2, which used a zirconium-nickel hydride pump. A hydrogen gas-gap CTSW and hydride pump/getter system is shown in Figure 8. In principle, the hydrogen gas-gap CTSW is nominally "off" (evacuated) until actuated "on" by heating a metal hydride getter, which evolves hydrogen and provides thermal conductance across the gap. Although the Swales hydrogen gas-gap CTSW demonstrated an "off" resistance greater than 1000 K/W with both a Zr-V-Fe metal getter and a turbopump and an "on" resistance less than 1.0 K/W, the SDCC design remains the preferable solution for near term space-borne applications, due to the higher anticipated reliability of a fluid-less/getter-less system. A potentially more reliable solution than the gas-gap/getter system is shown in Figure 9. Latching solenoid valves similar to those used on the HST Nicmos Cooling System, Nellis3, would be baselined for such a system. Further details on this system, the Swales gas-gap CTSW design, and Swales gas-gap CTSW test data may be found in Marland4.

Figure 7. Layout of the CRYOTSU Flight Experiment.

Figure 8. Hydrogen Gas-Gap CTSW and Hydride Pump System.

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Figure 9. Hydrogen Gas-Gap CTSW and Solenoid Valve/Helium Tank System.

Figure 10. Multi-Stage Cooler "Cross-Strapping" Concept.

CRYOCOOLER CROSS-STRAPPING

One argument against the use of CTSWs is the potential option of "cross-strapping" in multistage cooler systems. Figure 10 illustrates the concept. Cross-strapping, which is simply a thermal linking across the lower stages of redundant cryocoolers, uses the excess cooling capacity in the "on" cooler's lower stages to pump heat away from the "off" cooler's lower stages with little performance degradation of the "on" cooler's highest stage. The net result is that all the stages of die "off" cooler run considerably colder than they would if cross-strapping were not used. However, cross-strapping is not an option for single-stage systems, and most of the longlife coolers developed and flight qualified to date are, in fact, single-stage coolers. Thus, only if multi-stage coolers become more prevalent will cross-strapping reduce the need for CTSWs. Furthermore, even with cross-strapping, the thermal resistance of a properly designed CTSW is significantly greater than the thermal resistance between the two highest stages of the cooler. As a result, the CTSW increases the thermal isolation between the non-operating cryocooler and cryogenic component and reduces parasitic heat leaks. CTSW RELIABILITY

Based on the aforementioned discussions, it is evident that the system benefits of CTSWs are substantial. The decision to implement CTSWs should therefore be based on reliability. For

space applications, the CTSW should have a reliability of nearly 100%. Although the SDCC has


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not completed a rigorous flight qualification process, the anticipated SDCC reliability is expected to meet the rigorous standards of flight applications. The operation of the SDCC is

based on the reversible and highly reliable physical process of differential thermal contraction/expansion of stainless steel relative to beryllium. The simple three-piece design and successful thermal and mechanical cycling ground testing of the SDCC enforce the reliability expectations. The SDCC is now ready for flight qualification life cycle testing. ADVANCED SDCC DESIGN

By modifying the SDCC mounting scheme and large beryllium part, the SDCC mass and surface area-induced parasitics may be greatly reduced without effecting system operation or performance. Figure 11 illustrates the concept. The design shown in Figure 11 has a mass of just 50 grams with virtually no surface area-induced parasitics. As with the prototype SDCC design, direct mounting to the cryogenic component eliminates the need for a flexible link between the SDCC and cryogenic component, providing a significant improvement to the system

"on" performance and mass. SUMMARY

The primary objective of this paper is to describe the design, operation, and test results of the Swales differential CTE CTSW (or SDCC), which is designed to couple redundant cryocoolers to cryogenic components with minimal off-cooler parasitics. With its simple three-piece construction and highly-reliable, repeatable actuating mechanism of differential thermal contraction, the SDCC is intended to provide a high reliability CTSW to meet the DoD performance goals of 1000 K/W "off" resistance. In ground testing, the SDCC easily met these requirements, demonstrating an "off" resistance of 1400 K/W and an "on" resistance of 1.2-2.0 K/W over a cold end temperature range 25-50 K and a warm end temperature range of 230-300 K. In addition to the SDCC design and test results, related CTSW topics such as test methodology, gas-gap CTSWs and cross-strapping are addressed. Finally, an advanced SDCC design with a mass of less than 50 grams and virtually no area-induced parasitics is introduced. Cryogenic thermal switches are critical cooling and input power saving devices, which represent an important part of the Integrated Cryogenic Bus (ICB) initiative to incorporate new and enabling cryogenic technologies into space systems. Offering potential high reliability, simple construction, and unsurpassed performance, the SDCC provides intriguing promise for the near future implementation of cryogenic thermal switches in space applications. This paper is intended to provide the cryogenic space applications community with a SDCC status report.

Figure 11. Advanced SDCC Design.

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ACKNOWLEDGMENTS

The authors would like to acknowledge the Air Force Research Laboratory for providing the funding for this work. We would also like to acknowledge the efforts of Thorn Davis and B.J. Tomlinson of AFRL.

REFERENCES 1.

Bugby, D., Stouffer, C., Hagood, B., et. al, "Development and Testing of the CRYOTSU Flight Experiment," Space Technology and Applications International Forum (STAIF-99), M. El-Genk

editor, AIP Conference Proceedings No. 458, Albuquerque, NM, 1999, pp. 2-3. 2.

Johnson, D. and Wu, J., "Feasibility Demonstration of a Thermal Switch for Dual Temperature IR Focal Plane Cooling," Cryocoolers 9, Plenum Press, New York (1996).

3.

Nellis, G., Dolan, F., Swift, W., and Sixsmith, H., "Reverse Brayton Cooler for NICMOS," Cryocoolers 10, Kluwer Academic/Plenum Publishers, New York (1998).

4.

Marland, B., Bugby, D., Stouffer, C., "Development and Testing of Advanced Cryogenic Thermal Switch Concepts," Space Technology and Applications International Forum (STAIF-00), M. ElGenk editor, AIP Conference Proceedings No. 504, Albuquerque, NM, 2000.

Thermally Conductive Vibration Isolation System for Cryocoolers G. S. Willen Technology Applications, Inc. Boulder, CO, USA 80301

E. M. Flint CSA Engineering, Inc. Mountain View, CA, USA 94043

ABSTRACT

With the increasing demand for high temperature cryoelectronics and sensitive spaceborne infrared imaging and detection systems, reliable mechanical cryocoolers have become an enabling technology. These cryocoolers share two undesirable characteristics, vibration and heat generation. Vibration causes deflections in attached structures, adversely affecting sensor systems alignment and inducing spurious electrical signals. Heat generation increases cryocooler temperature, thereby reducing efficiency, reliability, and lifetime. Technology Applications, Inc. (TAI), with the support of CSA Engineering, Inc., developed a multi-axis Thermally Conductive Vibration Isolation System (TCVIS) for cryocoolers under a SBIR Phase II program for the Space Vehicles Directorate of AFRL (Contract No. F29601-97-C-0114). The vibration isolation goals were a multi-axis 50:1 vibration reduction at the cryocooler drive frequency, and a 10:1 reduction at two or more harmonics. INTRODUCTION

Split-Stirling cryocoolers generally use opposed compressors whose vibration can be significantly reduced by electronically adjusting their phase relationship. Back-to-back operation is not feasible for most cryocooler expanders and active axial vibration cancellation is employed; however, reduction of axial vibration has no effect on the lateral vibration. Therefore, this development study focused on multi-axis vibration isolation for cryocooler expanders. When isolating the expander vibration from the structure to which it is attached, the expander itself becomes effectively isolated from the structure. This poses a problem for in-space applications where expander heat removal depends upon conduction. Copper braid or aluminum foil straps are thermally inefficient, heavy, and difficult to integrate into the system. In addition they

are generally too stiff to decouple the expander vibration from the structure. Therefore, a lightweight, flexible, effective heat removal subsystem was integrated into the TCVIS design. The TCVIS, shown in Figure 1, fits between the cryocooler and detector dewar assembly. It incorporates six-axis vibration reduction that uses compact, efficient electromagnetic actuators, force sensors, and passive isolators. This system has demonstrated a hundred-fold reduction in

vibration at the fundamental frequency with significant reduction at multiple higher harmonics. Cryocoolers I I , edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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Figure 1. Thermally conductive vibration isolation system. The primary features incorporated into the design are ease of integration into cooling systems and adaptability to many different cryocoolers. The TCVIS was designed to accommodate a 1 -watt split-Stirling tactical cryocooler; a Hughes Model HM7050C-514 used in this program. These type of cryocoolers have demonstrated in excess of 6,000 hours operating lifetime, making them a low cost alternative for many ground, airborne, and space-based tactical infrared surveillance systems. For applications where vibration control and effective heat rejection are required, the TCVIS developed under this SBIR Phase II can provide those capabilities. DESCRIPTION

The major TCVIS assemblies and components are shown in Figure 2. The cryocooler is mounted to an aluminum flange that conducts heat from the expander to a high conductance, flexible graphite fiber thermal strap. The flange assembly consists of an aluminum upper flange that is attached to a stainless steel lower flange by a welded bellows assembly. The lower flange bolts to the detector dewar housing; the cryocooler expander is attached to the upper flange. The bellows connecting the two flanges forms a vacuum tight assembly and provides a low axial stiffness path that effectively decouples the expander vibration forces from the lower flange. The vibration isolation assembly consists of six electromagnetic actuator struts arranged in a Stewart configuration that provides the capability for both axial and radial vibration cancellation. The upper ends of the actuators are attached to three posts that are attached to the upper flange; the lower actuator ends are attached to the lower flange. Vibration forces generated by the expander are counteracted by the actuators, attenuating the forces transmitted to the lower flange. Heat Transport Subsystem Heat generated by gas compression is localized in the expander head. This heat is conductively transferred from the expander head to the upper aluminum flange via a conductive copper collar

THERMALLY CONDUCTIVE VIBRATION ISOLATION SYSTEM

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Figure 2. TCVIS expanded view showing the major subassemblies. attached to the expander. The heat is then conducted to the flexible graphite fiber thermal strap assembly that is bolted to the aluminum flange. The collar is made from oxygen free high conductivity (OFHC) copper, slips over the expander head, and is clamped with two cap screws to provide good thermal contact. It is bolted to the upper flange near the thermal strap-mounting interface.

This assembly is shown in Figure 3. The thermal strap assembly conducts the expander heat from the upper flange to the heat rejection interface. It is the most important heat transfer element between the cryocooler and the heat rejection radiator, as over half of the temperature drop occurs across the strap. In addition to effective heat rejection, the thermal strap has sufficient flexibility to effectively attenuate vibration, either from the TCVIS to the spacecraft structure, or from the spacecraft to the TCVIS.

Figure 3. Heat transport subsystem.

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Figure 4. Flexible graphite fiber thermal strap assembly. The low-stiffness, high-conductance requirement led to the development of a lightweight, flexible, high-thermal conductivity graphite fiber thermal strap. The strap assembly, shown in Figure 4, consists of a single row of 20-fiber bundles and two aluminum end fittings. It is made from high thermal conductivity K-l 100 graphite fibers attached to aluminum end fittings. These fibers have 2.75 times the thermal conductivity of copper and 1/4 the density; on a per unit mass basis, they are theoretically 11 times more weight efficient than copper, and over six times more weight efficient than aluminum. The theoretical maximum conductance is 0.45 W/K; the measured thermal strap conductance was 0.35 W/K, yielding a thermal efficiency of 78%. The weight of the strap is 26.5 g, approximately half the weight of a comparable aluminum foil strap.

Vibration Isolation Subsystem The six-axis vibration isolation system (VIS6) reduces both the axial and radial forces generated by the cryocooler. To provide multi-axis vibration isolation, the isolation system is arranged in a Stewart (i.e., hexapod) configuration, in which the six struts are oriented to provide control authority over all six principal degrees of rigid body motion. A stand-alone signal processor was developed using a feed-forward control algorithm to reduce the first five disturbance tones (the primary and four harmonics). The VIS6 consists of two subsystems: the transducer hexapod that contains the six struts and the electronics. The electronics consist of the load cell signal conditioning and the electronic support package. The VIS6 system schematic is shown in Figure 5. The six struts are arranged in symmetric pattern of three repeated pairs around the cryocooler expander; this arrangement is identified as the vibration assembly in Figure 2.

Each strut, shown in Figure 6, consists of a single electromagnetic proof mass actuator (PMA), a viscoelastic passive stage, a piezoelectric wafer load cell, and two axial end flexures; it weighs approximately 120 g. The electromagnetic PMA provides the counter force required by the control subsystem. A current supply power amplifier in the electronic support package (ESP) supports the actuator in each strut. Similarly, signal conditioning is provided to the wafer load cell of each strut by a dedicated operational amplifier. The average resonance of the actuators is 44 Hz, at which they can generate 0.47 N/V. At 200 Hz the actuators can generate 0.133 N/V. In the low frequency range (below resonance) the generated force was limited by the 23.3 mils end-stop limit. The filtered-x least mean square (FXLMS) algorithm was used as the basis for the active vibration control. FXLMS is a version of the least mean square algorithm in which the reference

signal is filtered through an estimate of the secondary path dynamics before it is used in the output and primary plant update.

SUBSYSTEM TEST RESULTS Tests were conducted on both the heat transport and vibration isolation subsystems to demonstrate performance and to characterize the individual subsystems. The thermal testing was limited to measuring thermal strap conductance. Extensive testing of the vibration isolation subsystem was conducted to fully characterized its operation and performance.


Figure 5. VIS6 system schematic.

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Figure 6. VIS6 strut configuration.

Thermal Strap Test Results

The thermal straps were characterized by measuring their overall thermal conductance. These tests were performed at the National Institute of Standards and Technology (NIST) in Boulder, CO and repeated at TAI; both sets of test data were in good agreement. The thermal conductance was measured and compared with theoretical predictions to determine overall strap conductance and thermal efficiency. Each strap was tested at a minimum of five different power levels to determine if the conductance was dependent on power. A flexible strap with 20-fiber bundles was fabricated and tested; refer to Figure 4. Its measured conductance was 0.35 W/K; this value represents an average of five data points taken at different heater power settings. The theoretical maximum conductance is 0.45 W/K; therefore, the thermal efficiency of the flexible strap is 78%. Different power settings were used to verify the test measurement consistency and to assess whether increasing the amount of heat transferred (heat flux) affected the conductance. The conductance variation with power and the average temperature drop across the strap for each series of tests is shown in Figure 7. These results were typical for all thermal straps tested and showed the conductance to be independent of heat flux. The data also provide a check on the “goodness” of the conductance testing. For a constant conductance, the should approach zero as the heater power approaches zero. As shown in Figure 7, the linear extrapolation of the nearly intersects the origin.

Figure 7. Thermal conductance test results for a 20-fiber bundle strap.

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Figure 8. Comparison of open and closed loop behavior in the time domain for all six load cells at 4

Vrms cryocooler drive voltage. To verify these results, two additional thermal straps were fabricated and tested. Both were the same design, but they were fabricated with 18- and 22-fiber bundles. Their measured conductances were 0.34 and 0.36 W/K, respectively, which is consistent with the 0.35 W/K conductance of the 20-fiber bundle thermal strap. Vibration Isolation Test Results The vibration isolation subsystem performance was characterized over a cryocooler drive voltage range of 2-to-10 Vrms (10 Vrms corresponds to full power). Figure 8 illustrates a representative example of the vibration subsystem performance. It corresponds to the load cell response of strut 4 at a cryocooler drive voltage of 4 Vrms. Using the strut load cell locations and orientations, the axial and radial loads generated by the cryocooler and transmitted through the load cells were calculated. On average, the primary frequency was reduced 40.8 dB (100:1) in the axial direction and between 31.5 to 37.1 dB (37:1 to 70:1) in the two radial directions. Peak reductions of 45.6 dB (190:1) occurred at an 8 Vrms cryocooler drive voltage for the axial direction and 42.5 dB (130:1) at 4 Vrms for the radial x-direction. The next four tones that were controlled were reduced on average 21.8 dB (12:1) in the axial direction and between 15 to 19.4 dB (5:1 to 9:1) in the two radial directions. Overall performance results stayed relatively consistent versus the cryocooler drive level, as illustrated by Figure 9, which shows the axial and radial vibration force reductions for three cryocooler drive voltages. The radial reduction is an average of the two transverse radial components. The maximum axial and radial reductions of –40 and –35 dB, respectively, occur at the cryocooler

Figure 9. Vibration isolation performance for 8, 9, 10 Vrms cryocooler drive voltage.


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Figure 10. TCVIS test setup with thermocouple locations.

drive frequency (dark bars); the –20 to –25 dB reduction is an average of the next four harmonics (light bars). These results confirm that the VIS6 subsystem met or exceeded all vibration isolation

goals and requirements. INTEGRATED SYSTEM TEST

To demonstrate the TCVIS can meet operational requirements in a space environment, an integrated system test was conducted at NIST in Boulder, CO. The TCVIS was mounted onto a base plate with the heat sink attached to a liquid cooled cold plate, as shown in Figure 10. The test setup was attached to a heavy ballast plate using two fiberglass channels for thermal isolation. It was then placed inside a vacuum chamber where a diffusion pump maintained the vacuum below torr and a recirculating chiller provided the cooling for the heat rejection interface over a –20°C to +20°C temperature range. The criteria for determining successful TCVIS operation during the Integrated System Test was defined as 1) demonstrating a reduction in vibration forces comparable to those demonstrated during the subsystem level testing and 2) maintaining the cryocooler expander temperature below 70°C under the worst-case conditions of maximum cryocooler power and heat rejection interface temperature. Exact duplication of the extensive vibration subsystem testing was not expected because of the TCVIS mounting in the NIST vacuum chamber and the many additional vibrational modes associated with the vacuum chamber, diffusion pump operation, chiller flow transfer lines, and general disturbances throughout the laboratory. This test demonstrated the TCVIS met its target of multi-axis, vibration isolation, and that the thermal transport subsystem maintained the cryocooler expander within its operating temperature range. Vibration force reduction in the simulated space environment was demonstrated by reading the broadband vibration signal from the load cells, and viewing each strut load cell output on a digital oscilloscope. Thermal performance was determined by measuring critical component temperatures and the temperature distribution along the TCVIS conductive path. The broadband vibration reduction for the integrated test was not as great as that measured during the controlled subsystem tests but was generally in good agreement with the subsystem level test. The average steady state expander temperature, for a heat rejection interface temperature of

20°C, ranged from 60°C to 68°C as the power was varied from 80 to 100%. As expected, the temperatures increased as the power increased. Over the full power range, the thermal transport

subsystem was able to keep the expander temperature below 70°C, within the safe operating range of tactical cryocoolers.

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Figure 11. VRS1 mounted to the cryocooler expander.

TECHNOLOGY SPIN-OFFS

Two technology spin-offs, a single-axis vibration reduction system (VRS1) and a flexible thermal strap assembly, were developed during this program. The single-axis vibration reduction system is a counter-force based system that uses an electromagnetic proof mass actuator (PMA), acceleration based error signal, and a feedforward control algorithm implemented on a stand-alone processor. The VRS1, shown in Figure 11, is simple, compact, and attaches to the cryocooler or expander to facilitate ease of integration. It consists of a transducer unit, attachment fitting, control electronics package and associated cabling. The transducer unit (actuator and sensor) is bolted to a circular aluminum mounting ring that slips over, and is clamped to, the cryocooler expander head. When assembled, the actuator end cap sits on a ring of material separating it from the expander head, limiting heat conduction into the actuator. A flat-heat-rejection interface is machined into the aluminum mounting ring with three tapped holes for easy attachment of a thermal strap. The VRS1 reduces 11 disturbance tones, the primary and the first 10 higher order harmonics. Figure 12 shows representative results from the measurement of the VRS1 performance when mounted on the VIS6 hexapod with the cryocooler driven at 6 Vrms. The primary and second harmonic disturbances were reduced by over two orders of magnitude (42.5 and 40.5 dB). Over the

Figure 12. VRS1 performance measured by the accelerometer error sensor, 6 Vrms drive voltage.


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Figure 13. Graphite fiber thermal strap assembly.

bandwidth of operation (550 Hz), the measured axial rms acceleration disturbance levels were reduced, on average, by more than an order of magnitude (26 dB). At 10 Vrms, the primary and second harmonic disturbances were reduced by 18 and 28 dB, respectively; the higher harmonics

demonstrated an average force reduction of 28 dB. The decrease in performance at the primary and second harmonic is most likely due to VRS1 stroke limitations. This system in its entirety adds an additional 1.3 kg, of which 88% is associated with the electronics, cabling, and power conditioning. The system requires 9.6 W in standby mode, and 13.0 W when in the control mode. It is easy to retrofit onto existing systems and only requires access to the cryocooler drive signal and 18-36 V DC power.

Flexible Graphite Thermal Strap Assembly A lightweight, flexible, high-conductance, K1100 graphite fiber thermal strap (GFTS) assembly was developed. A 28.5-cm long GFTS with a conductance of 0.20 W/K has demonstrated a 78% weight savings over a solid copper (OFHC) bar of the same conductance. An equivalent

aluminum (1100F) rod would have a diameter of 0.75 in and weigh 2.7 times more than the GFTS. These copper and aluminum conductor weights assume solid rods; flexible braid or foil conductors would weigh substantially more. The GFTS configuration tested is pictured in Figure 13; it consists of two rows of K1100 carbon fiber bundles and two aluminum end fittings. Each row is made up of 20 fiber bundles; each fiber bundle consists of 20,000 individual fibers. Three 28.5-cm straps were fabricated with 1, 2, and 3 rows, respectively. Their theoretical and measured conductances along with their thermal efficiencies are summarized in Table 1. The thermal efficiency of the three straps shows a significant decrease when increasing the number of rows. The difference in thermal efficiency appears to be linear with the number of rows, about 8.5% for each row. To see if this trend is valid for four rows, a four-row engineering prototype strap, 24.8-cm long, with 88 total fiber bundles was tested. Its theoretical and measured conductances were 0.61 and 0.31 W/K, respectively. This yielded a thermal efficiency of 51%, indicating that the fourth row does not offer any significant advantage.

Therefore, the bulk of the inefficiency appears to be related to the end fitting conductance losses. SUMMARY AND CONCLUSIONS A TCVIS was developed that fits between the cryocooler and the sensor housing dewar assembly. It integrates both a thermal transport and a six-axis vibration cancellation subsystem into a

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compact, lightweight assembly. This system was designed to accommodate a 1-watt split-Stirling tactical cryocooler expander; however, it is adaptable to many different cryocoolers. At the cryocooler drive frequency, the TCVIS demonstrated more than 40 dB reduction in the axial direction and between 31 and 37 dB in the two radial directions; the next four harmonics were reduced an average of 22 dB. The flexible thermal strap successfully limited vibration transmission and maintained the cryocooler expander well within its recommended operating temperature. Using the technologies developed in this program, two innovative technology spin-offs were developed: a single-axis vibration cancellation system for tactical cryocooler expanders and a flexible thermal strap assembly. The single-axis vibration reduction system reduced 11 disturbance tones. The primary and second harmonics, the dominant axial disturbances, were reduced by over two orders of magnitude; the higher harmonics were reduced by more than an order of magnitude. The graphite fiber thermal strap assemblies demonstrated thermal efficiencies up to 92%, yielding a 65 to 80 percent weight savings over aluminum or copper straps for the same conductance. These straps have sufficient flexibility to accommodate installation and alignment tolerances, relative structural movements due to vibration and thermal contractions, and limited bending. ACKNOWEDGEMENTS The authors wish to acknowledge Mr. Michael Evert, Mr. Patrick Flannery, and Dr. Eric Anderson of CSA Engineering, Inc., for their significant contribution in the development of the vibration isolation subsystem; Mr. Edward Myers and Ms. Jennifer Lock of Technology Applications, Inc., for their contributions to the system design, analysis, and testing; and Dr. Dino Sciulli, AFRL Technical Project Officer, for his assistance and guidance throughout the project.

Advanced Cryogenic Integration and Cooling Technology for Space-Based Long Term Cryogen Storage B. J. Tomlinson and T. M. Davis Air Force Research Laboratory

Kirtland AFB, NM 87117 J. D. Ledbetter

Mission Research Corporation Albuquerque, NM 87110

ABSTRACT

The Air Force Research Laboratory (AFRL), Space Vehicles Directorate has been the lead Department of Defense (DoD) agency for the development of low capacity cryogenic refrigerators and integration technologies for space applications since the late 1980s. High

capacity cryocoolers and long term (>20 years) on orbit propellant storage are potentially enabling technology for future High Energy Laser (HEL) space systems, orbital transfer vehicles, and on orbit propellant depots. Cryogenic applications in space based systems requiring long term cryogen storage includes: significant cooling requirements for subcritical cryogens, cryocooler redundancy issues, on orbit cryogen transfer from vehicle to vehicle, large shield cooling, long term gas and liquid cryogen storage, large distributed cooling surfaces, cryogenic system integration issues, and significant spacecraft system penalties due to mass and input power. AFRL has pursued low capacity cryocooler concepts including reverse Brayton cycle, single and multiple stage Stirling cycle, advanced Joule-Thomson cycle, and Pulse Tube (Stirling cycle variant) designs. The cryocooler technology spans a wide range of cooling temperatures (from ~10 Kelvin to 150 Kelvin) and heat loads (up to 10 Watts at 95 Kelvin). Additionally, AFRL has pursued advanced cryogenic integration technology including cryogenic thermal switches, cryogenic heat transport, thermal storage, and cryogenic integration schemes to reduce system mass and input power penalties. Current cryogenic integration and cryocooler development programs address the negative impacts of the cryogenic system on optical space systems including: induced line of sight vibration, longevity, power consumption, mass, thermal transport, thermal storage, and thermal switching. However, the cryogenic cooling requirements for future Air Force systems may require large capacity cryogenic cooling, extremely mass and power efficient mechanical refrigerators, and significant improvements in long term on orbit cryogen storage. The technical efforts at AFRL concentrate on exploratory and advanced development programs that focus on Cryocoolers 11, edited by R.G. Ross, Jr. Kluwer Academic/Plenum Publishers, 2001

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the development of technology from concept and breadboard engineering models to protoflight models that are geared to experimental characterization and technology transition for flight demonstrations and, potentially, operational programs. INTRODUCTION

For the last decade, developments in cryocooler technology have focused in on low capacity heat lift in the 35 Kelvin to 150 Kelvin temperature range with recent developments in cooling at 10 Kelvin. Essential elements of the technology have evolved to establish a technology base for long life cryocoolers designed to meet reliability, mass, vibration, and thermodynamic performance requirements. In addition, cryogenic integration technology has also focused on low capacity thermal issues in the cryogenic system. Evolving mission requirements for infrared, multi, and hyperspectral focal plane and optics cooling are still driving the technology development to lighter, more efficient and reliable cryocoolers.1 Future Air Force and Department of Defence (DoD) systems including High Energy Laser (HEL) space systems, orbital transfer vehicles, and on orbit propellant depots all have requirements for the long-term storage of cryogens, also known as zero-boil off systems, in a liquid or gaseous state for use as propellant or chemical reactants. In addition to the Air Force and DoD system concepts, the National Aeronautic and Space Administration (NASA) has been exploring technology for human and robotic space exploration. The long-term cryogen storage requirements cover numerous issues including increased cooling capacity refrigerators,

cryogenic system redundancy, cooling of large distributed surfaces, cryogenic system integration, and significant system power and mass reduction.

This problem is not a new one. The idea of long-term on-orbit cryogen storage has been considered in the literature for at least three decades. One of the largest obstacles to implementing the past system concepts was the lack of reliable, sufficiently efficient, and low mass cryogenic cooling technologies. Current technology development for sensor and optics cooling has provided a potential bridge for the development of feasible long-term cryogen storage systems.2 The requirements for the long-term storage of cryogens are still emerging and could change many times before the completion and implementation of a flight system. However, there are many common attributes of the possible system configurations that can be covered by an organized technology development program. Past AFRL technology development programs have focused on lower heat lift capacity, long life technology and significant system issues such as reduction of induced vibration, mass, and input power. However, current AFRL programs provide a starting point for technology development to meet the advanced requirements. The following sections provide an overview of current AFRL programs for cryocooler and cryogenic integration development and outlines potential plans to meet the emerging requirements. POTENTIAL REQUIREMENTS

There are potential requirements on the horizon that will require expansion of past developments to increase the reliable lifetime, reduce the mass, significantly increase the capacity, and increase the power efficiency of cryocooler and cryogenic integration technology. System concepts for Air Force HEL systems and orbital transfer vehicles are in their infancy and still evolving. Specific cryogenic capabilities are very system design dependent, but there are common themes for the range of technology requirements. Table 1 provides some examples of potential cooling requirements of large tanks and shields for on-orbit reactant storage for a conceptual space based HEL system. However, the cooling temperatures and loads are only part of the problem. The issues of how this system is integrated, potential redundancy, the large cooling surfaces on the shields and tanks, and the system impact of cooler mass and input power are significant and especially system dependent. Decisions on development of large capacity coolers, gangs of small capacity coolers, single stage coolers, or multiple stage coolers depend on the particular cryogenic system schemes that will be pursued.

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As an example, redundant multistage cryocoolers could mitigate the need for cryogenic thermal switches to offset the parasitic heat load of the “off” redundant cryocooler(s) by crossstraping the stages and intercepting the heat at the more efficient warmer stages of the cooler to prevent most of the heat load from reaching the less efficient colder stages of the “on” cooler. The Aerospace Corporation has examined several potential scenarios for cryogenic storage of reactants for space based HEL systems. This includes trades on cryogenic gas versus liquid systems, large versus small coolers, small versus large tanks, and other system level attributes. An additional consideration was the comparison of system cooling performance by utilizing hypothetical multistage cryocoolers that were based on extensions of current technology. The results of these initial studies have shown that there is a wide range of system impacts depending on the specific system design. As expected, the tank heat leaks were shown to have a large impact on required cryocooler cooling power and with lower efficiency coolers the input power required was staggering for even a small system. However, with more effective tank thermal isolation and multilayer insulation, and with an efficient cryocooler, the input power and mass penalties become manageable.3 An additional finding was the comparison of systems employing single stage cryocoolers or multistage coolers for a conceptual 100 Kelvin tank of gaseous cryogen. The results from Aerospace show that the multistage systems have significant system benefits over single stage cooler systems.4 Cryogen storage in space and for the exploration of extraterrestrial bodies is also a technology need for the National Aeronautical and Space Administration (NASA) Mars exploration programs. Human and robotic missions to Mars will require cryogen liquefaction, storage, and transfer for propellant and breathable atmosphere during the long trip to Mars and for storage on the surface for use at a later time. Kittel, et. al., has examined the problem of cryogen storage and liquefaction cooling for conceptual Mars exploration and although the mission lifetime requirements are on the order of 6 months to five years, the technology required to meet these mission requirements will apply to future AF and DoD missions. Potential cooler requirements for the Mars missions are shown in Table 2 and are a mixed combination of various missions and durations up to 1.4 years.5 As system concepts for AF and DoD missions evolve, the requirements for advanced cryogenic integration and cooling technology for space based long-term cryogen storage will change. However, there are current development activities that are addressing many of the grass roots issues associated with the cooling system requirements for these advanced systems. As is the case on many development programs that span decades, the component level technology maturity and capabilities are major drivers in the system design. Careful development and investment into cooler and cryogenic integration technology will allow for the timely maturity of

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the technology through both ground proof-of-concept demonstrations, prototype hardware development, and flight qualification through short or long duration space missions and Advanced Concept Technology Demonstrations (ACTD). CURRENT DEVELOPMENT ACTIVITY The Refrigerators

When examining the potential system concepts that are on the horizon it becomes evident that some of the requirements will need the capability for large capacity cooling. As the system concepts evolve, the extent of the increase in capacity over the current state of the art designs will become clearer. There is the possibility of utilizing a single large cryocooler to produce all of the necessary cooling for a given requirement, or sharing the load between two or more smaller coolers. This is a system design, integration, and reliability issue and will be considered during the evolution of the future concepts. Although significant improvements in the state-of-the-art have been realized, there are still development issues to be resolved, especially in cooling ranging from 10 to 50 Kelvin. However, the developments such as improved non-contacting “Oxford” flexure bearing compressors, miniature gas bearing turbine compressors and expanders, pre-cooled and lubrication-less Joule-Thomson systems, and marked improvements in pulse tube technology have greatly expanded the use of cryocooler technology in a broad range of space applications and are baselined for many near-term and future missions as “enabling” technology. This is a reflection of the increased mission need for cooled sensors and optics, but is also a reflection of the increasing maturity of cryocooler and cryogenic integration technology. AFRL is currently developing higher capacity cryocoolers to meet the Space Based Infrared System Low (SBIRS Low) technology requirements for the Engineering Manufacturing Development phase of the program development. It was identified at the beginning of the AFRL technology development that there was a need for highly efficient, low mass cryocoolers for cooling the gimbaled fore optics of a conceptual space based system. The need arose due to the large cooling load that would be required to cool the necessary optics, the need to minimize the waste heat on gimbal by increasing efficiency (due to limited radiator area on gimbal), and to reduce the mass as much as possible due to the large amplification of the mass penalty on gimbal. The requirement was determined to be 10 Watts of cooling at 95 Kelvin, with minimal mass and a specific power of at least 10 Watt input power / Watt cooling or lower. The two programs currently underway at AFRL are the TRW 95K High Efficiency Cryocooler (TRW HEC) and the Raytheon 95K High Efficiency Cryocooler (Raytheon HEC). Both development programs are leading to technologies that will provide the framework for the follow-on generations of high capacity machines. The TRW 95K High Efficiency Cooler is the latest development in pulse tube technology. The cooler is a vast improvement in pulse tube cryocooler technology due to a combination and culmination of various technology developments. The cooler is designed to lift 10 W of heat at 95 K with an overall cooler specific power of 10 W/W.TRW has transitioned the “Oxford” linear compressor technology developed under IR&D and under the IMAS cryocooler development program for the Jet Propulsion Laboratory. The compressor developed for the 95K program has greatly reduced mass over the state of the art and will have 6 cubic centimeters of swept volume and a compressor efficiency of 83%. The mass of the whole cooler will only be 4.5 kg. The first cooler under this program will be completed and delivered to AFRL in November 2000. NASA Ames Research Center is utilizing this AFRL / TRW development to procure an additional unit under this contract for demonstration of the Mars liquefaction requirements discussed in the requirements section. NASA will utilize the cooler for ground demonstrations of the in-situ liquefaction and storage of gases simulating the Martian atmosphere in preparation for near and mid-term follow-on flight programs to Mars. In its present form, a number of these coolers would be required to meet the 80 K and below cooling requirements described in Table 1. However, it is reasonable to believe that with a

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development program, the technology used to create the compact and power efficient compressor can be scaled up to much larger swept volumes. TRW has past experience in large capacity pulse tube development and, with some development, can scale up their recent developments to accommodate the higher capacity cooling requirements. Although this may sound like a

“simple” extension of current technology, it is not. There are significant issues in the design of a large capacity pulse tube cryocooler. Issues such as non-linear scaling of internal design features, heat rejection interfaces (external and internal), and regenerators may pose significant research and development problems for cooler designers. The Raytheon 95K High Efficiency Cryocooler (Raytheon HEC) is the latest extension of the successful technology development under the Raytheon Protoflight Spacecraft Cryocooler program, delivered to the Air Force and is currently in endurance evaluation at AFRL. The Raytheon HEC cooler is poised to be a very versatile for sensor and optics cooling in addition to many other different potential applications. The program requirements for this cooler are

identical to the TRW HEC cooler program and is being designed for 10 W at 95 K, with a specific power of

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