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NBS TECHNICAL NOTE 1049

U.S. DEPARTMENT OF COMMERCBNational Bureau of Standards
NATIONAL BUREAU OF STANDARDS

The National Bureau of Standards' was established by an act of Congress on March 3, 1901. The Bureau's overall goal is to strengthen and advance the Nation's science and technology

and facilitate their effective application for public benefit. To this end, the Bureau conducts

research and provides: (1) a basis for the Nation's physical measurement system, (2) scientific and technological services for industry and government, (3) a technical basis for equity in trade, and (4) technical services to promote public safety. The Bureau's technical work is per- formed by the National Measurement Laboratory, the National Engineering Laboratory, and

the Institute for Computer Sciences and Technology.

THE NATIONAL MEASUREMENT LABORATORY provides the national system ol physical and chemical and materials measurement; coordinates the system with measurement

systems of other nations and furnishes essential services leading to accurate and uniform

physical and chemical measurement throughout the Nation's scientific community, industry,

and commerce; conducts materials research leading to improved methods of measurement,

standards, and data on the properties of materials needed by industry, commerce, educational

institutions, and Government; provides advisory and research services to other Government

agencies; develops, produces, and distributes Standard Reference Materials; and provides

calibration services. The Laboratory consists of the following centers:

Absolute Physical Quantities' — Radiation Research — Thermodynamics and

Molecular Science — Analytical Chemistry — Materials Science. THE NATIONAL ENGINEERING LABORATORY provides technology and technical ser- vices to the public and private sectors to address national needs and to solve national problems; conducts research in engineering and applied science in support of these efforts; builds and maintains competence in the necessary disciplines required to carry out this research and technical service; develops engineering data and measurement capabilities;

provides engineering measurement traceability services; develops test methods and proposes

engineering standards and code changes; develops and proposes new engineering practices;

and develops and improves mechanisms to transfer results of its research to the ultimate user. The Laboratory consists of the following centers:

Applied Mathematics — Electronics and Electrical Engineering 2 — Mechanical

Engineering and Process Technology2 — Building Technology — Fire Research —

Consumer Product Technology — Field Methods.

THE INSTITUTE FOR COMPUTER SCIENCES AND TECHNOLOGY conducts

research and provides scientific and technical services to aid Federal agencies in the selection, acquisition, application, and use of computer technology to improve effectiveness and

economy in Government operations in accordance with Public Law 89-306 (40 U.S.C. 759), relevant Executive Orders, and other directives; carries out this mission by managing the Federal Information Processing Standards Program, developing Federal ADP standards

guidelines, and managing Federal participation in ADP voluntary standardization activities; provides scientific and technological advisory services and assistance to Federal agencies; and

provides the technical foundation for computer-related policies of the Federal Government.

The Institute consists of the following centers:

Programming Science and Technology — Computer Systems Engineering.

'Headquarters and Laboratories at Gaithersburg, MD, unless otherwise noted; mailing address Washington, DC 20234.

2 Some divisions within the center are located at Boulder. CO 80303.
A Study of Design Principles for Refrigerators

for Low-Power Cryoelectronic Devices

J.E. Zimmerman

D.B. Sullivan

Electromagnetic Technology Division

National Engineering Laboratory

National Bureau of Standards

Boulder, Colorado 80303

\ as /

U.S. DEPARTMENT OF COMMERCE, Malcolm Baldrige, Secretary

NATIONAL BUREAU OF STANDARDS. Ernest Ambler. Director

Issued January 1982
National Bureau of Standards Technical Note 1049

Nat. Bur. Stand. (U.S.), Tech. Note 1049, 114 pages (January 1982)

CODEN:NBTNAE

U.S. GOVERNMENT PRINTING OFFICE

WASHINGTON: 1982
CONTENTS

Page

1. Introduction 1

2. Background

2.1 Split Stirling-Cycle 2

2.2 Twin-Piston Stirling Cycle 4

3. Summary of Refrigerator Development

3.1 Split Stirling Cycle 4

3.1.1 Materials 4

3.1.2 Construction 5

3.1.3 Operation 6

3.2 Twin-Piston Stirling Refrigerator 8

3.3 Compressor Design 8

3.3.1 '0' Ring Sealed Piston with Conventional Crosshead 8

3.3.2 Force Balanced Compressor 9

3.3.3 Contamination Free Compressor 9

3.4 Temperature Control. . 10

4. Cryocooler Optimization

4.1 General Approach 11

4.2 The Calculation 13

4.3 Results 14

5. Summary and Conclusions 17

6. References 17

7. Appendix 19

m

A Study of Design Principles for Refrigerators

For Low-Power Cryoelectronic Devices

J. E. Zimmerman and D. B. Sullivan

National Bureau of Standards

Electromagnetic Technology Division

Boulder, Colorado 80303

This report summarizes a five-year effort at NBS which has been directed toward the

development of low-power cryocoolers suited to the support of superconducting instruments.

The report deals with a variety of aspects of construction and operation of refrigerators as well as with a model which allows one to optimize the design for minimum drive power. The

publications generated by the program are included as an appendix.

Key words: cryocooler; cryogenics; low temperature; refrigerator; Stirling cycle; superconducting devices.

1. Introduction

This report describes a five-year effort at NBS to develop small, low-power Stirling crycooolers

which are suited to the support of highly sensitive superconducting instrumentation, particularly

magnetometers and magnetic gradiometers based on the Superconducting Quantum Interference Device

(SQUID). The latter part of this first section describes the environmental requirements for such

instrumentation. In Section 2 we provide a brief discussion of the operation of the cryocoolers and in

Section 3 summarize the refrigerator development work. Section 4 describes a model which has been

developed to optimize cryocooler design and finally the publications generated by the program are appended for completeness.

The new class of cryocoolers considered in this report were first described at two conferences in the summer of 1976 [1, 2]. The machine described in these papers achieved a temperature of 13 K with

three stages. The addition of a fourth stage reduced the bottom-end temperature to less than 8.5 K and

a niobium SQUID magnetometer was operated with the cooler [3]. Analysis of the machine's operation was reported subsequently [4, 5]. The operation of the cooler is limited by the internal thermal

resistance of the plastic regenerator components, a situation much different than that for all other

Stirling machines. In 1979, a low-pressure, twin-piston single stage cooler was built. This cooler

succeeded in cooling from temperatures in the 8 to 12 K region to a bottom end temperature of 3 to 4 K

[6]. This stage could conceivably be added to one of the higher-pressure (2 to 4 bars), multistage

machines to achieve helium temperature operation. Finally we refer to a sequence of papers [7, 8, 9,

10,] which review and update the progress on this refrigeration work. The last of these papers

describes a 5-stage refrigerator which achieves a bottom end temperature of less than 7 K with only 20

W of mechanical power input. This machine has been fitted with a small diagonal -component SQUID

gradiometer which shows, in a preliminary way, that SQUIDs and refrigerators can be successfully

integrated. For demonstration purposes this gradiometer-cooler system has produced a magnetocardiogram

of modest clarity [11]. This combination was also used to investigate the level and source of magnetic

interference generated by the cryocooler. Some preliminary work has also been done on a pulsed
technique for measuring the thermal characteristics (thermal conductivity and specific heat) of the

plastic regenerator materials used in these machines [12]. This method is useful because the usual

measurement methods are cumbersome due to the very long thermal response times for these materials.

The environmental requirements for a SQUID magnetometer are extreme and, as far as we know, no attempt to operate a SQUID with a commercial cryocooler has been made. The expected vibration and

magnetic noise levels are so large that the SQUID output would be meaningless [13]. This is not surprising since the SQUID has a sensitivity of the order of 10 times the earth's magnetic field.

Thus, even ignoring the direct magnetic interference from moving parts in the refrigerator, a motion of

4 only one microradian in the earth's field causes an interference output which is 10 times the

sensitivity. From these figures one readily appreciates the need for vibration reduction. Low speed

and low-power operation can produce such reduction. In fact, we have taken the view that optimization

of refrigerator design for this purpose is really a minimization of the refrigerator power input. The

assumption is that the net interference will scale with the size and power requirement of the drive

machinery. The calculations in Section 4 are based on this design philosophy. Obviously, one must

also avoid the use of magnetic components and even good conductors (since motion of a conductor in a field produces field variation through induced eddy currents). Even ostensibly non-ferro-magnetic

materials have susceptibilities and remanent moments which can be troublesome. Thus, material

selection is important. It is not essential, however, to eliminate all interference, since the major

interference component is synchronous with the drive system of the refrigerator. Synchronous filtering

can thus be used to reduce interference, but it is certainly best to start with a low level of

interference because such filtering can never be complete.

Three final points should be made. First, if refrigerators are to become important for the

support of superconducting devices, then it is essential that their complexity and cost be not much

greater than that of the devices they support. This places some sort of bound on practical solutions

to the problem. Second, there is a possibility that high-T devices can alter the situation. That is,

the cooling requirements are significantly affected by use of devices that might operate at say 10 or

15 K. Fabrication technology for these devices is not highly developed, but good devices have been

reported by several laboratories, perhaps most notably the Naval Research Laboratory. Where such

devices could be used, the size and cost of the cryocooler would be reduced significantly [14].

Finally, the feature of most commercial cryocoolers which contributes to the interference is an unnecessarily large refrigeration capacity and drive power (of the order of kilowatts). By matching

the crycoooler capacity to the inherently low refrigeration requirements of a SQUID magnetometer, a large reduction in drive power and direct magnetic interference from the drive train can be realized.

Cost can be reduced and portability enhanced at the same time. This is really just a restatement of

the design philosophy mentioned in the previous paragraph.

2. Background

2.1 Split Stirling-Cycle

The operation of a split Stirling-cycle refrigerator can be understood through study of figure 1. Gas moves freely by the displacer exchanging heat with the displacer and walls as it moves between the

hot end (T,) and cold end (T~). Thus, the pressure throughout the system is nearly the same at any

instant in time. Work W = fPdV may be done on the system as the piston moves back and forth, but no work is done by or on the displacer. When the displacer is at its lowest position, the piston is moved

to the left compressing the working fluid (helium gas) and producing heat Q-, , which is rejected at T-,.
Then the displacer is moved to the top of the cylinder so that gas is displaced from top to bottom.

Assuming that the machine has been operating for some time and the bottom temperature T« is less than

T\ , then the gas gives up heat Q, to the displacer and wall in the annular gap and arrives at the

I a bottom at a temperature Tp. The piston then moves to the right producing a heat flow Q^ into the gas at T~. Finally, the gas is moved upward as the displacer is returned to its bottom position and a heat

Q. is added as the gas temperature goes from T~ back to T,. This completes one cycle of operation.

<

ciT 2 REGENERATOR

GAP

PISTON

Absorbed

REGENERATOR

HEAT CAPACITY

(Q a = Qb for Perfect Gas)

Figure 1. Single-stage, split-cycle Stirling refrigerator with phase diagram to show the idealized

thermodynamic cycle.

The process of heat exchange between the gas and the materials in the annular gap is called

regeneration. For efficient regeneration the heat capacity of the solid must be much greater than that

of the working fluid. This is usually easy to achieve above 10 or 20 K for helium, but the specific

heat of solids relative to helium drops rapidly below this temperature leading to large inefficiency.

This, along with the non-ideal behavior of helium near its boiling point generally limits the bottom

end temperature in the 5 to 7 K range. At the sacrifice of cooling capacity one can reduce the helium

pressure to enhance the heat capacity ratio. This is the basis for the 3.1 K temperature achieved by

the twin-piston refrigerator [6].

The Stirling-cycle refrigerator described above is readily adapted to multistage operation. A single stage machine such as shown schematically in figure 1 was found to cool to less than 50 K from

room temperature [1]. Figure 2 shows the simple modification which converts the device to multistage

operation. At each step in the displacer there is an annular space at which cooling takes place. The

successively lower steps (stages) are equivalent to the addition of another refrigerator which rejects

its heat to the step above. Concentric radiation shields from each stage enclose all lower stages so that radiation is intercepted and pumped at intermediate temperatures where refrigeration efficiency is high. Further discussion of the split Stirling-cycle refrigerator is contained in references 1 and 10.
2.2 Twin- Piston Stirling Cycle

We have used the twin-piston Stirling-cycle for only one low-pressure experiment and this work is not currently continuing, so we will not dwell on this particular configuration. Suffice it to say that two pistons operating from the same crankshaft can accomplish the task of shuttling gas back and

forth through a regenerator as well as providing the required compression and expansion necessary to approximate the cycle shown in figure 1. This mode of operation proved useful for a low-pressure

machine which achieved a temperature of 3.1 K [6, 8].

3. Summary of Refrigerator Development

3.1 Split Stirling Cycle

In this Section we summarize the construction details and performance characteristics of four

coolers which are listed in 3.1.4 below. The four machines use one, three, four, and five stages

achieving temperatures less than 50 K, 13 K, 8.5 K and 7 K, respectively.

3.1.1 Materials

The four coolers all use nylon as a displacer with spun-glass epoxy (designated G-10 by the

National Electronic Manufacturers Association) as the cylinder material. The radiation shields shown

in figure 2 are fabricated of thin-wall aluminum, but could equally well be composite fiberglass

structures with strips of foil running along their length. The requirement is a reasonable axial

thermal conductivity with little or no circumferential eddy currents. The composite shields have been

used successfully in many fiberglass dewar systems and should work well here. The aluminum was used

here to speed construction. Besides the vacuum insulation, aluminized mylar is used to reduce

radiation losses. We have used this material without the usual fibrous spacer.

Plastic Piston

in Cylinder

Helium

Gas

Radiation Shields

Cold End

Squid

Figure 2. Multistage split-cycle Stirling refrigerator.
3.1.2 Construction

The glass-epoxy (G-10) tubes are purchased from manufacturers stock in a form which is immediately

usable. That is, very little machine work is required on these tubes. The tubes are laid up on very

precise mandrels so the interior finish and dimensions are well controlled. The exteriors of the tubes

are ground and are often not concentric with the inner surface. This is an important fact which must

be considered in the assembly of multistep machines. To construct a machine such as shown in figure 3,

the outer end of each section is machined so that the outside part which slips inside the adjacent

cylinder is concentric with its interior surface. The connections between cylinders and to the

aluminum flange and end cap are all made with epoxy. We have found that it is best to machine these

parts so that they mate with an interference fit. Thus, the epoxy is used primarily as a sealant and

bond rather than a filler. These joints have been highly reliable.

Dlsplecer

Pueh Rod Aluminum

Flange ^ Solid Nylon

> Stepped Dleplacer

(In lowest position)

Aluminum Spacer

(where required)

Glaee-Epoxy Composite

Cylinder Walla

Thin Aluminum

Radiation Shields

Alumlnlzed Plastic (10-20 layera In vacuum apace)

Aluminum End Cap

Cold Space For Device

^Vacuum Wall

Figure 3. Details of a 5-stage, split-cycle Stirling refrigerator.

The drawing is split in two parts to show finer detail.

Dimensions are in millimeters.
The nylon displacer is formed from stock rods which are pinned together to form a multistage

machine. Each rod is individually fitted to its appropriate cylinder prior to assembly of the displacer or the stepped cylinder. In general, the nylon stock is slightly too large and must be machined

to fit the cylinder. These displacer sections are carefully machined to fit the cylinder with almost

no radial clearance, except for a few cm at the large end where about 0.1 mm radial clearance is provided. During cooldown, differential expansion between the nylon and fiberglass produces a one- percent radial clearance in the lower portions of the refrigerator. This fitting method is simple and

minimizes the regenerator gap volume, which is important for achieving lowest temperature.

The differential expansion in the axial direction must also be considered. The lengths of the

displacer sections are selected as as to match the lengths of the corresponding sections of the

cylinder when the system is at its lowest operating temperature. This requires some foreknowledge of

the temperatures of the different stages and an estimate of the differential expansion as a function of

temperature. After initial cooldown the displacer is adjusted so that (ideally) all steps reduce their

respective volumes to zero at the bottom of the stroke.

Figure 3 provides dimensional information (in millimeters) on the 5-stage cooler. The upper end

of each radiation shield is slit so that it can be slipped over the appropriate step on the cylinder

and secured with 20 to 40 turns of thin nylon string (unwaxed dental floss is ideal). The bottom ends

of these aluminum tubes are closed with aluminum foil or aluminized mylar which is also secured with

dental floss.

Mylar superinsulation is used between each shield and between the outer shield and the vacuum

wall. We have found that this insulation can be omitted with little effect in the space between the

cold end and the first shield.

All instrumentation wiring (thermometer leads, thermocouple wire and rf leads to SQUIDs) are thermally anchored at each stage. In one instance, we failed to fully anchor one pair of leads at the next to bottom stage and found significant deterioration in the bottom end performance. Thermal

anchoring is achieved by securing the leads tightly with dental floss and then coating the area with a suitable varnish.

3.1.3 Operation

The displacer push rod in figure 3 is fitted with an adjusting screw so that the bottom limit of

the displacer motion can be altered as the system cools down. The range of this adjustment is about

one percent of the displacer length.

After purging the system of air (either simple purging with helium gas or pumping the air out),

the system is filled with helium to between 2 and 4 bars pressure. The piston and displacer are initially set to 90 phase difference and the system is put into operation. From this point on there

are three adjustments which can be made. These are the phase, the pressure and the displacer spacing

from the cylinder bottom. The latter is the simplest, since the adjustment can be made by noting the

point where the displacer hits bottom and then backing off slightly from that position. The

adjustment of pressure and phase for lowest bottom end temperature is generally different from that for

fastest cooldown and the selection of proper parameters for either optimization is tedious since the

system responds very slowly. It is quite easy to be misled in these adjustments. For example, in optimizing for lowest temperature one often finds that a small adjustment in phase or pressure results

in a downward trend in temperature, but this trend is only a transient and the final steady state can,

in fact, be higher than the initial temperature. This behavior follows from the fact that these

parameters affect every stage of the refrigerator and the upper stages respond slowly because of their

higher total heat capacities. For the 5-stage machine, we have found that the optimum pressure is between 2 and 2.5 bars (gauge pressure) and the phase is closer to 45° than 90°.
With the sliding '0' ring seals on the displacer push rod and in the compressor (see Section 3.3),

there is a very gradual buildup of air contamination in the system. This freezes into the lower parts

of the displacer and cylinder and gradually causes friction and sticking of the displacer with a resultant increase in temperature. The system must be warmed up and purged of these contaminants

before recooling. Presently this problem occurs every one to three weeks depending upon the quality of

the sliding seals. With further engineering work this problem should be eliminated.

A second problem results from the fact that helium diffuses through the spun glass-epoxy

material. Thus, the vacuum space must be continuously pumped to maintain the integrity of the

insulation. The pump can be turned off for only short periods (15 min. to an hour) before the

temperature starts to rise. In principle, this problem could be resolved by laminating a thin metal

foil into the glass-epoxy cylinder to block the diffusion.

3.1.4 Discussion and Summary of Split Stirling Refrigerators

In table I we summarize the characteristics of the four machines which have been built and tested.

The one-stage cooler was really operated with a compressed gas cylinder and valves, that is, in a Gifford-McMahon cycle, although it could easily have been converted to Stirling cycle. The bottom end

temperature would probably have been about the same. Measurement and modeling of shuttle heat loss [4,

5] indicate that such loss can be reduced by shortening the displacer stroke. This is the reason for

the shorter stroke in the 4 and 5 stage machines.

TABLE 1

CHARACTERISTICS OF FOUR MULTISTAGE CRY0C00LERS

Lowest Temperature

Compressor Di splacement (cc)

Di am. /Length - Stage 1

(both in mm) Stage 2

Stage 3

Stage 4

Stage 5

Stroke (mm)

Temperature - Stage 1

1- Stage

50 K

*

9.6/300

3-Stage 4- Stage

13 K 8.5 K

38 38

19/245 28/150

9.45/143 19/120

4.7/144 9.5/100

4.7/150

12.7 7

120 K 180 K

12.7

50 K

Stage 2 40 K

Stage 3 13 K

Stage 4 8.5 K

Stage 5

*This machine was operated with valves by compressed gas (i.e., Gifford-McMahon cycle).

5- Stage

7 K

90

38/140

29/130

19/120

9.5/83

4.8/133

7

167 K

70 K

27 K

15.5 K

7 K
The addition of more and more stages leads one closer to a continuously tapered displacer, and we

now believe that this might in fact be the next thing to do. The construction will require some custom

fabrication of the displacer and cylinder, but we suspect this may really be simpler than the present

construction which requires a lot of careful fitting and joining of stock tubes and rods. Another

point to make is that a tapered displacer model can really offer a fair approximation to a stepped

displacer machine. In section 4 we consider such a model and find that it describes the stepped

displacer results quite well.

3.2 Twin-Piston Stirling Refrigerator

This development is described rather completely in the paper by Zimmerman and Sullivan [6] with

some additional discussion in reference [8]. Thus we omit extensive description of the device here. We will emphasize only a few points which are not made in those papers. The idea behind this machine was to gain lower temperature in a single low- temperature stage. By

reducing the helium pressure to values below atmospheric, one lowers the boiling point of the liquid

and thus maintains nearly ideal gas behavior until this lower temperature is approached. More

importantly, the ratio of regenerator heat capacity to gas heat capacity is enhanced and improves the

regenerative action.

Operating from upper temperatures of 8 to 12 K the machine achieved bottom end temperatures of 3.1 to 4 K as the average pressure was varied from 0.26 to 0.79 bars. In all cases, a small amount of gas was liquified in the bottom of the cooler, and the temperature achieved was just the boiling point of

helium at the average operating pressure, e.g. 3.1 K at 0.26 bars. To our knowledge, this is the

first Stirling refrigerator in which such liquefaction has been demonstrated.

The twin-piston machine described in reference [6] used epoxy fiberglass for all the lower end

components, so there was no need for concern about differential expansion. The two pistons were set to

a 90 phase difference and no further phase adjustments were made. Such adjustment would have been

difficult because the crankshaft and connecting rods were sealed into a helium-filled enclosure.

This device could be coupled to one of the split Stirling machines to achieve helium temperature

operation in one system, but it must be operated as a completely separate machine because of the low

operating pressure. The complication of combining two refrigerators has caused us to concentrate on further improvement of the split Stirling-cycle refrigerator in the hope that this machine will prove

suitable for most applications.

3.3 Compressor Design

We report here on the details of construction of three different compressors, a simple '0' ring

sealed piston with conventional crosshead, a balanced force device, and a ceramic oil-free system.

3.3.1 '0' Ring Sealed Piston with Conventional Crosshead

The simplest compressor was constructed with a nylon piston having one or two '0' rings for seals

with a felt pad loaded with silicon grease in an adjacent groove to provide lubrication. A sketch of

such an arrangement is found in reference [1]. This piston was driven at approximately 1 stroke per

second by a motor and crosshead made up of an old commercial freon compressor. The 38 cc compressor

had a piston diameter of 35.6 mm and a stroke of 38 mm. The 90 cc compressor used a stroke of 51 mm with a piston diameter of 48 mm.
3.3.2 Force Balanced Compressor

There are two problems with the simple compressor. First, the force needed for compression

requires a rather large motor. Much of the work done in compressing the gas can be recovered during

the expansion phase. Thus it is possible to balance this force with other conservative forces so as to reduce the torque needed to operate the compressor. The first and simplest modification is to keep

helium gas on the back side of the piston at a pressure which is the average value of the system

pressure. This leaves the system at equilibrium with the crankshaft 90 ahead or behind full

compression which reduces the required torque. Further reduction can be accomplished through the use of a spring and cam as shown in figure 4. An added benefit is that the piston '0' rings are not directly exposed to the atmosphere. Air leakage is only through the seal around the piston rod.

TO CRANKSHAFT

BALL BEARINGS

IN CHANNELS

self-centering^ USHAFT COUPLING

ADJUSTED TO

LEAK (EQUALIZES AVERAGE PRESSURE)

TO DISPLACER

Figure 4. Sketch of mechanism used to balance the compressor. The spring and cam are adjusted and

shaped so that the torque at the crankshaft is nearly constant.

3.3.3 Contamination Free Compressor

Ultimately one should address the problem of a design which has no lubricated seals and no possibility for buildup of air contamination. While the latter has been our more immediate problem, it

is clear that oil-free operation is essential for continuous, long-period use. Over a long period of

time oil vapor would build up deposits in the cold portion of the refrigerator with resultant

deterioration in performance. There are several approaches to the design of a contamination-free

compressor, and while this is not the major thrust of our effort, we describe here some preliminary

experience with one concept.

The compressor shown in figure 5 uses very closely spaced ceramic (alumina) cylinders to effect a seal during compression. The radial clearance of 4 micrometers allows very little pressure loss during

compression and the back side of the piston is sealed with a metal bellows to complete the gas con- tainment. The time constant for pressure equalization across the piston is of the order of one minute.

Thus, leakage past the piston is negligible at the operating speed of one stroke per second. This

compressor was operated approximately 60 hours at which point the bellows failed. Subsequently, we replaced the bellows with an '0' ring seal. This system was run non-stop for over 10 months. The

results of this test indicate some initial wear but little or no wear after about 1000 hours. The wear

is inferred from measurement of leak rate past the cylinder (see figure 6). There is some discrepancy

in the measurements beyond 3000 hours probably due to miniscule accumulation of contaminants on the

surfaces. The measurements at 7100 hours showed a small increase in leak rate when the surfaces were

cleaned with toluene.
Bellows Ceramic

^)) )))W) wwmnF

Figure 5. Details of design of an oil-free compressor. The clearance between ceramic components is 4 micrometers keeping gas leakage at a negligible level during compression. The compressor was actually

operated vertically.

0.3

a

co01

«

A

Euo

T

gas —

0.34 MPa

(gauge)

Ul

1 liter cylinder

Schematic of Leak Rate

Measurement

water

CERAMIC COMPRESSOR - Helium Working Fluid

Bore X Stroke - 5 cm X 5 cm

Speed - 1 Hz

Temperature - 2 1 C

Pressure Excursion - 0.2 to 0.7 MPa (absolute)

1 1 1

1 2 3 4 5 6

TIME lOOOHours

Figure 6. Wear measurement for the ceramic compressor. Wear is inferred from change in leak rate with

time. The experimental setup is shown in the insert at top. BDC and TDC stand for bottom dead center

and top dead center, respectively.

3.4 Temperature Control

In abandoning the use of liquid helium in favor of a miniature cryocooler, it is necessary to take measures to stabilize the temperature of the cold end of the cryocooler. We have observed that small

variations in operating parameters (such as ambient temperature, speed, pressure, impurity buildup,

etc.) cause significant slow variations in cold-end temperature, of the order of 0.1 K or more. Also, we see a temperature oscillation at the cryocooler frequency (1 Hz). This oscillation is probably

10
quite large in the expansion chamber itself, but may be quite small at the superconducting device,

depending on the thermal resistance between it and the expansion chamber, and on the thermal mass of

the device and associated hardware. In the four-stage cryocooler, the amplitude of the oscillation was about 2 mK at 8.5 K, as measured by a carbon resistance thermometer cemented into a brass block screwed

to the bottom of a helium gas thermometer bulb clamped to the cold end of the cryocooler (see figure

7). The amplitude measured by the gas thermometer was about 5 mK. It is to be expected that the

amplitude of temperature oscillation should be larger, the closer the thermometer is to the expansion

space. It will also, of course, be a function of the response time of the thermometer.

In order to stabilize and control the temperature of the brass block (which simulates a SQUID with

mounting hardware) shown in figure 7, we assembled the simple circuit shown in figure 8 to amplify the

dc-signal from the carbon resistance thermometer and feed it back to the 10,000-ohm resistance heater.

Thus if the control resistance R in the bridge is set lower than the thermometer resistance R-r with

feedback loop open, the closed-loop output of the amplifier will provide power to the heater R,, to raise the temperature of the brass block until the bridge is balanced. Initial tests of this control

system showed that slow changes in temperature become imperceptible (no more than a fraction of a mK in

a two-hour period). Because of the rather slow thermal response, the simple circuit did not significantly reduce the 2 mK oscillation at 1 Hz. We intend to work on electronic regulation of the

oscillatory component in the future.

Another method for stabilizing the temperature is to increase the heat capacity of the object

attached to the cold end. This might best be done with helium gas in a small volume between the cold

end and the SQUID. The situation would be like that of figure 7. We found that the heat capacity of

this system could easily be increased by a factor of 10 or more by filling the 1 cc volume to only 4 or

5 bars of pressure. This results in a 10-fold reduction in the oscillatory component of temperature

variation. This also reduces the drift rate significantly.

This latter method of temperature stabilization can readily be developed further. The concept is analagous to electronic filtering with resistor-capacitor combinations. The heat capacity of an object

is the analog of electrical capacitance and thermal resistance is the analog of electrical resistance.

Using this analog, one realizes that thermal resistance should be introduced between the refrigerator

and gas reservoir to improve the filtering.

4. Cryocooler Optimization

4. 1 General Approach

The general design philosophy for these cryocoolers involves one fairly simple assumption; to minimize refrigerator interference one should operate with the smallest possible drive power. In other

words, the larger the drive machinery, the greater will be the magnetic and vibrational

interference for any SQUID or other system supported by the refrigerator. Thus, when we say optimum

design, we mean minimization of the power requirement for the system. We have selected a tapered displacer Stirling cooler for this calculation with the following

thoughts in mind. First, we can cover the entire temperature range from just above liquid helium

temperature to room temperature with a single calculation. Second, the results will provide design

guidance, even for stepped displacer machines. That is, the result of the calculation is an optimum

shape and size for the displacer and this can be used to estimate the lengths and diameters of the

stages in an optimum stepped displacer machine. Finally, we would like to consider the construction of

a tapered displacer system in the future.

The idea in this calculation is to express the refrigeration and losses as time-averaged functions

of the parameters of the system (heat capacity, thermal conductivity, position along the displacer,

11
stroke, frequency, etc.). The losses are balanced against the refrigeration with the functional form

of the taper (cylindrical symmetry) finally determined by using a variational method to minimize ^he

power required to run the system.

NYLON DISPLACER

EPOXY-GLASS CYLINDER

EXPANSION SPACE

ALUMINUM END PLUG

GAS THERMOMETER

BULB (BRASS)

HEATER (10 kO)

CARBON RESISTANCE THERMOMETER

(220 NOMINAL)

BRASS BLOCK

Figure 7. Bottom-end details of the 4-stage cooler with electronic temperature control

the bulb adds significant thermal mass and thus adds to temperature stability.

Helium gas in

BRASS BLOCK

5M

DIFFERENTIAL

AMPLIFIER

-±r COMPENSATED

INTEGRATOR

Figure 8. Simple temperature control bridge circuit. The amplifier delivers current to the heater so

as to match R T to R . c

12
4.2 The Calculation

In this calculation we consider three loss terms; conduction, radiation, and shuttle heat loss.

These are the dominant losses for the temperature range of about 15 to 300 K. Below 15 K one should

include the enthalpy deficit and regenerator inefficiency. These terms are more complex because they

include the mass flow of the working fluid. Thus, as a first step we consider the limited problem

which nonetheless results in useful design information in the temperature range above 15 K. Despite

the potential for some error below 15 K, we have extended the calculations to 10 K. The net heat leak can be expressed as

^ ~ ^cond ^rad ^shuttle

= 7tr 2 k(dT/dx) + 2eoT

3Af(x) (dT/dx) + S

2 r(nnkC/2) 1/2 (dT/dx) , (1)

where k is thermal conductivity and C is specific heat, both of which are functions of x. r is the

radius, S is the stroke, A is the surface area of the radiation shields, n is the operating frequency,

and e and a are constants relating to radiation. f(x) is a function which describes the distribution

of insulation along the refrigerator, x is the position along the displacer and the radius r is taken

as a general function of x, so that the taper profile is as yet unspecified. The third term, shuttle

heat loss is taken from reference [4].

The thermal conductivity and specific heat of the displacer and cylinder are functions of

temperature and must be explicitly included in the calculation. From a combination of sources including reference [4], we have generated the following power series representations of thermal

conductivity, k(T), and specific heat, C(T) for the G-10 material in the range from 5 K to 300 K.

k(T) = 9.17 x 10" 2 + 3.19 x 10" 3 T - 3.67 x 10 _6

T

2

, (2)

C(T) = - 1.49 x 104

+ 3.75 x 10

3T + 3.35 x 10

] T2

- 9.21 x 10 _2

T

3

, (3)

•3 where k(T) is in units of W/K-m and C(T) is in J/K-m . The diameter at the bottom end of the displacer is set by equating the refrigeration at the bottom

end with the heat leak into the same point, that is,

2 Q(r

o

) = 7rr o Sn P,/n(P

u /P£

) (4)

13
where r is the bottom end radius and P and P. refer respectively to the upper and lower pressures.

The expression on the left is the sum of the heat losses (eq. 1) evaluated at r = r

Q

(or equivalently,

at x = 0). The refrigeration term is an ideal one; it should be reduced by about one half for the

real sinusoidal drive.

The ideal work required by the refrigerator is given by

W=/((Tamb

/T - 1) (dQ/dx) dx , (5)

where the integral is over the length of the machine. The first term inside the integral is the Carnot

coefficient and T , is the ambient (room) temperature.

Equation 4 can be solved for r as a function of T (the bottom temperature) and (dT/dx) ~. Thus

if we select an arbitrary function, T(x), to describe the temperature profile, we can calculate a value

for r and then, having differentiated Q with respect to x, we can calculate the net work (eq 5) for

this temperature profile. Equation 4 can also be used to find r as a function of x; since the heat

flow at a point, x, is intercepted by the refrigeration below that point to the displacer length.

The results described are for a specific T(x) and thus do not represent any sort of optimum, but we do have all of the ingredients for a variational determination of the optimum T(x) and thus displacer shape. We select a power series for T(x),

T(x) = T

Q

+ T

1 X + T2 X

2 + T

3

X

3 + . . . . , (6)

where the sum of the coefficients, T , is equal to T

b

- T

h

. . if x is normalized to the displacer

length. An arbitrary set of coefficients (usually only a linear term) is chosen and after calculating

the work, the coefficients are varied to find if this increases or decreases the work and only changes

which decrease the net work are accepted. By successive iterations of this process for all of the coe- fficients we converge upon the best possible temperature profile and displacer shape. The absolute

length of the displacer is an input parameter.

4.3 Results

Figure 9 shows the results of this computer generated solution for a 50 centimeter long displacer

which is comparable to that used in our 4 and 5 stage machines. The stroke is 5 mm, the cylinder wall

thickness is 3 mm, and the frequency is 1 Hz. The 4 radiation shields are taken to have an average

diameter of 10 cm and 40 layers of aluminized mylar are distributed between these shields. The upper

and lower pressures are 6 and 3 bars, respectively. These parameters are consistent with the values

for the 4 and 5 stage coolers described in section 3. The net work (ideal) required in this

refrigerator calculation (10 K bottom end temperature is 0.564 watt with 0.211 watt going to conduction

loss, 0.214 watt to shuttle heat loss, 0.082 watt to radiation loss, and 0.057 watt required to cool

the bottom end. The bottom end term is primarily shuttle and conduction heat flow which is proportional to the temperature gradient and the cross-sectional area at the bottom end. It would be zero if the wall thickness and displacer radius could be tapered to zero, but this is obviously

impractical. Our choice of a constant 3 mm wall thickness over the whole length was arbitrary, but the

14
variational method would work for any functional dependence of wall thickness on axial position. Since

our refrigerators require tens of watts of input power, we conclude that most of that power goes into

compressor losses (non-isothermal operation), motor losses, and friction.

The distribution of refrigeration to the different loss mechanisms is noteworthy. With modest

effort, radiation loss is relatively small. Conduction and shuttle heat loss are comparable. The

quantitative agreement between theory and experiment is also shown in figure 9. The stepped displacer

radius should be compared with the tapered displacer radius. The agreement is remarkable if three

facts are considered. First, the comparison is between a refrigerator with a stepped displacer and a model with a continuously tapered one. Second, in the experiments the parameters were adjusted to achieve the lowest temperature which may not be obtained for the set of parameters which yield the

minimum drive power. Third, the temperature achieved was 8.5 K rather than 10 K. The only serious discrepancy between analysis and experiment is in the mechanical power required

to run the compressor. We have always estimated that the order of 10 watts or more is required,

whereas the analysis (see above) gives the order of 0.6 watt. To resolve this discrepancy, we recently

measured the pressure-volume (P-V) diagram of the 4-stage machine (see table 1) with the average

pressure adjusted to give a steady-state bottom-end temperature of 10 K. The area of the P-V diagram

multiplied by the operating speed gave 2 watts as the mechanical power actually performed on the

working fluid. Correcting for non- isothermal compression reduced this to a net power of 1.3 watts. In comparing this with the analytical optimum of 0.6 watts, it should be noted again that a stepped

displacer is not expected to be as efficient as a continuously tapered displacer, and furthermore the

experimental machine had about twice the diameter of the analytical model.

300

~ 200

Ul

~~l I

1

Stroke

Frequency

Wall Thickness

Upper Pressure

Lower Pressure

Radiation Shields

5mm

1Hz

3mm

1 x 105 Pa

5x104 Pa

4

10 20 30 40

POSITION ALONG DISPLACER (cm)

50

Figure 9. Temperature and displacer radius as a function of position along the radius. This is the

optimum shape described by the model for the set of parameters listed above. The stepped displacer for

the 4-stage machine is superimposed for comparison with the radius of the tapered displacer.

15
With this calculation as a reference, we successively varied the stroke, the length, and the

frequency to see whether better values could be predicted. The results are shown in figure 10. While

the length of a half meter was selected somewhat arbitrarily, the stroke and frequency were varied

experimentally to arrive at the values given by the circles in figure 10. It is interesting to note

how closely the calculation predicts the best set of values for these parameters. The optimum stroke

of * 3.5 mm is close to the value of 5 mm used in the refrigerators, and while the minimum power occurs

for a displacer length of one meter, the half meter value which is more practical does not produce too dramatic a rise in system power. As for the frequency, it is clear that one should operate at the

lower end of the flat portion since the inclusion of regenerator efficiency will probably cause the

right end of the curve to be lifted. The gradual rise in power requirement with displacer length

reflects the increase in radiation loss which is itself quite small.

I I I I

g1.0 - y^ -

oc

UJ2 0.6 - N. y^

0.2 i i i i

oc

LUo

o.

4 6

STROKE (mm)

0.4 0.8 1.2 1.6 DISPLACER LENGTH (m)

1.6 2.4

FREQUENCY (Hz)

4.0

Figure 10. Power requirement as a function of stroke, speed, and displacer length. Each of the 3 parameters is varied separately with the remaining parameters held at the values given in figure 9.

16
One can also account for excess dissipation by some apparatus at the bottom of the refrigerator.

For the reference conditions (parameters) described above we note a linear increase of require drive

power, 35.7 mW, for each mW of refrigeration at 10 K for power dissipations of up to 20 mW. There is,

of course, a commensurate increase in the diameter of the bottom end of the displacer.

5. Summary and Conclusions

It now appears that this new class of low-power Stirling cryocoolers can be compatible with

superconducting sensors, at least for modest sensitivity. Several engineering problems remain to be

solved. First, the compressor needs to be designed for dry operation to eliminate any contact between

the helium gas and oil. This will require some effort, although the low speed and low working pressure

offer possibilities for simple solutions. The problem of helium diffusion through the displacer

cylinder should be eliminated with some sort of diffusion barrier. Finally, considerably more work

should be done to isolate and identify magnetic interference mechanisms. This could permit extension

of the devices to more demanding applications.

All things considered, we feel that the analytical results agree remarkably well with the

experimental measurements, insofar as valid comparisons can be made. Thus, the tapered displacer model seems to provide reasonable guidance for refrigerator design. It would be especially useful to extend

the calculations to include regenerator inefficiency and enthalpy deficit. The present calculations

suggest that a linear taper (cone) will suffice for temperatures down to 10 K, but we suspect that the extra losses below this temperature will require a more complex taper. In any case, it should prove

useful to build a continuously tapered system, since success with this concept could lead to improved

performance and simplified construction.

6. References

[I] Zimmerman, J. E. , Radebaugh, R. , and Siegwarth, J. D. , in "Superconducting Quantum Interferences

Devices and their Applications," (H. D. Hahlbohm and H. Lubbig, eds.), Berlin, Walter de Gruyter

(1977), pps. 287-296.

[2] Zimmerman, J. E. , Radebaugh, R. , and Siegwarth, J. D. , "Refrigeration for Small Superconducting

Devices," Conf. German Refrigeration Society, Munich (October 1976).

[3] Zimmerman, J. E. and Radebaugh, R. , in "Applications of Closed-Cycle Cryocoolers to Small Superconducting Devices," NBS Special Publication SP-508, J. E. Zimmerman and T. M. Flyn, eds.), Superintendent of Documents, U. S. Government Printing Office, Washington, D. C. 20234 (1978). (Also

available from the authors), pp. 59-65.

[4] See ref. 3 pps. 67-73.

[5] Radebaugh, R. in "XVth International Congress of Refrigeration," Venice, Italy, Sept. 1979 (In

press).

[6] Zimmerman, J. E. and Sullivan, D. B. , Cyogenics ]9, pp. 170-171, (1979).

[7] Zimmerman, J. E. in "Future Trends in Superconductive Electronics," (B. S. Deaver, Jr., C. M. Falco, J. Harris, and S. W. Wolf, eds.) AIP Conf- Proceedings Number 44, New York, American Institute

of Physics, 1978, pps. 412-420.

[8] Sullivan, D. B. and Zimmerman, J. E. , Int. J. of Refrig. 2, pps. 211-213, (1979).

[9] Zimmerman, J. E. in "Proceedings of the SQUID Geophysics Workshop," Los Alamos, NM, June 1980. (In

press).

[10] Zimmerman, J. E. in "SQUID 80: Superconducting Quantum Interference Devices and their Applications" (H. D. Hahlbohm and H. Lubbig, eds.) Berlin, Walter de Gruyter (1980), pps. 424-443.

[II] Sullivan, D. B. , Zimmerman, J. E. , and Ives, J. T. in "Refrigeration for Cryogenic Sensors and

Electronic Systems," NBS Special Publication SP-607, (J. E. Zimmerman, D. B. Sullivan, and S. E. McCarthy, eds.), Superintendent of Documents, U. S Government Printing Office, Washington, D. C. 20234

(1981). (Also available from the authors), pps. 186-194.

17
[12] Zimmerman, J. E. , Sullivan, D. B. , Kautz, R. L. , and Hobbs, R. D. , See ref. 11, pps. 173-177.

[13] Cox, J. E. and Wolf, S. A., see ref. 3, pps. 123-129.

[14] Nisenoff, M. , see ref. 11, pps. 195-209.

18
7. Appendix - Publications generated by the Cryocooler Program. The original page numbers have been

deleted so as to provide continuous numbering in this volume. Original page numbers are included in the reference at the beginning top of each publication.

Page

1. Possible Cryocooler for SQUID Magnetometers 20

2. Refrigeration for Small Superconducting Devices 30

3. Operation of a SQUID in a Very Low-Power Cryocooler 38

4. Shuttle Heat Transfer in Plastic Displacers at Low Speeds 45

5. Analysis of Regenerator Inefficiency for Stirling-Cycle Refrigerators with

Plastic Displacers 52

6. A Milliwatt Stirling Cryocooler for Temperature below 4 K 58

7. Cryocoolers for Superconductive Electronics 60

8. Very Low-Power Stirling Cryocoolers Using Plastic and Composite Materials 69

9. Cryocoolers for Geophysical SQUID Magnetometers 72

10. Cryogenics for SQUIDs 75

11. Operation of a Practical SQUID Gradiometer in a Low-Power Stirling Cryocooler 96

12. Measurement of Thermal Properties of Cryocooler Materials 105

19
"Superconducting Quantum Interference Devices and their Applications," (H. D. Hahlbohm and H. Liibbig, eds.) Berlin, Walter de Gruyter (1977), pps. 287-296.

POSSIBLE CRYOCOOLERS FOR SQUID MAGNETOMETERS*

J. E. Zimmerman, R. Radebaugh and J. D. Siegwarth

Cryogenics Division, National Bureau of Standards,

Boulder, Colorado 80302 USA.

ABSTRACT

In a study to determine if a cyclic cryogenic refrigerator is

suitable for cooling SQUID magnetometers and similar instruments, we have used a specially designed Stirling machine with

a three-stage displacer to achieve and maintain a temperature

of 13 K. The displacer and cylinder were made of low-susceptibility, nonconducting plastic to minimize magnetic interference.

For the same reason, and also to minimize mechanical noise,

the machine operates at low displacement and low speed, and it uses only about 10 W of mechanical power, requiring 50 W input

to a typical small electric motor. Since the temperature

achieved is within the range of NbN and Nb^Sn SQUIDs, it can,

in principle, be used to cool a SQUID. The estimated magnetic

signal from the reciprocating displacer may not seriously

affect the magnetometer sensitivity, but the signal due to the

cyclic pressure-induced geometric distortion of the SQUID

mounting in the earth's field may be very difficult to eliminate,

INTRODUCTION

In this paper we address a subject that is always implicit,

but not usually expressed in detail, in any application of

superconductivity, namely, the cryostat. For small devices,

such as SQUIDs, the universal practice is to use liquidhelium cryostats with a variety of shapes, sizes, and special

features. (1) It need hardly be pointed out that the cost and

inconvenience of procuring and handling liquid helium, the

attendant restrictions on how the instrument can be mounted

and handled, and the inherent intermittent nature of the

operation, all contribute to the obvious disadvantage of

superconducting devices relative to competing devices operating

at or near ambient temperature.

The alternative to liquid-helium cooling is to use a closedcycle cryogenic refrigerator (hereafter referred to as a cryocooler) . The use of a cryocooler replaces the disadvantageous features noted above with a whole set of new ones. Limited reliability is perhaps the most often mentioned;

others are maintenance requirements, cost, and, as we shall

see, magnetic and mechanical interference due to moving parts.

With regard to reliability, there seem to be two schools of

thought, corresponding roughly with the high reliability ex- pounded by cryocooler manufacturers on the one hand, and the

Contribution of NBS, not subject to copyright.

20
low reliability expected by potential users on the other. (2)

The fact is that very few cryocoolers are in use today to

operate small superconducting devices, so there is no substantial body of experience regarding reliability or anything

else.

There are several intermediate solutions to the cryogenic

cooling problem. One of these is a liquid-helium cryostat

incorporating one or more radiation shields maintained at

rather low temperatures by an auxiliary cryocooler, which may

greatly reduce the helium evaporation rate. (3) Such a system

combines some of the best, and some of the worst, features of

both worlds. One of the good features is that it permits one

to get practical experience in the use of cryocoolers without

being fully committed, in the sense that if the cryocooler

turns out to be no good one still has a cryostat. Another

useful feature is that the cryocooler can be turned off for

periods of several hours of critical measurements, where the

magnetic and mechanical noise of the cryocooler might be

intolerable, with only a slight increase in liquid-helium

consumption.

There are some other types of cryostats which may be useful

for particular applications. For example, a cascaded JouleThomson system operating from cylinders of compressed gases

may be suitable for intermittent short-term operation with

quick cooldown. -It is beyond the scope of this paper to

discuss all possible cryostatic systems, even by way of introduction, and we will hereafter restrict the discussion to

the use of liquid-helium and closed-cycle refrigerators

(cryocoolers) using helium gas as the working fluid.

In fact, there is only one known type of practical, small,

self-contained system which can maintain superconducting

temperatures in a 300 K environment. This is the type using

helium gas as a working fluid in a cycle involving isothermal

compression at room temperature and isothermal expansion, or a combination of isothermal expansion and other processes, at one or more low temperatures. The particular point to be

emphasized here is that compression and expansion require

moving parts, which may be very difficult to operate in a way

that will not interfere with the operation of highly sensitive

magnetometers

.

We have chosen the Stirling machine as an analytical and

experimental model to investigate the problems associated with

operating SQUID magnetometers with a small closed-cycle heliumgas cryocooler. It is a very simple machine, shown schematically in Figure 1, which has only two essential moving parts,

a piston and a displacer. (4) The displacer fits loosely in

its cylinder so that gas can move freely past it. Thus the

pressure P is nearly the same throughout the total volume V of

21
PISTON

T 2 REGENERATOR

GAP

Q, Rejected

Absorbed

REGENERATOR

HEAT CAPACITY

(Qa= Qb for Perfect Gas)

Figure 1 Schematic representation of operation of split

Stirling machine and idealized thermodynamic cycle

on temperature-entropy diagram.

the system. Work W = /PdV may be done on the system as the

piston moves back and forth, but no work is done by or on the

displacer since displacer motion does not change the volume.

Referring to Fig. 1, when the displacer is in its lowest

position, the piston is moved to the left, compressing the

working fluid and producing a heat of compression Qi which is

rejected at ambient temperature T^, typically 300 K. Next,

the displacer is moved to the top of its cy liner, so the

working fluid moves (is "displaced") from the top to the

bottom of the cylinder. If we may assume that the machine is

already in steady-state operation, with the bottom end at a low temperature T2, and a stable temperature gradient already

established along the displacer, then the fluid being displaced gives up heat Q to the displacer and cylinder walls

along the annular gap, and arrives at the bottom end at temperature T2. Then the piston is moved to the right, expanding

the fluid and producing a flow of heat Q2 into the fluid at

temperature T

2 at the bottom of the cylinder. Finally, the

displacer is moved to its lowest posj-tion, displacing the

remainder of the fluid back to the top of the cylinder and

completing the cycle. The fluid picks up heat Qb from the

walls of the annular gap as it travels up from T2 to T,

.

In steady state operation it is required that Q = Q^,

otherwise the temperature gradient along the displacer and

22
cylinder will change with time. With real gases, such as helium at 20 K and below, the requirement Qa = Q^ is incompatible with the assumption of isothermal heat exchange Qi and

Q2 at T^ and T2

. (i.e., The enthalpy change Qa or Q^ between

two given temperatures T^ and T 2 depends on pressure, so the

requirement Qa = Qu means that the expansion cannot be strictly isothermal. This is a limitation, or at least an analytical complication, in the operation of the Stirling cycle.

A more serious limitation, which we will have more to say

about, is that the heat capacity of the walls at the cold end

becomes insufficient, relative to that of the fluid, to

efficiently provide the necessary heat exchange Q and Q, . Considerable research has been devoted to finding materials

with favorable properties (large heat capacity and thermal

conductivity) to use in regenerative heat exchangers at low

temperatures. For use with SQUID magnetometers we would like

to impose additional restrictions on the material, namely,

that it also have favorable properties relative to generation

of electromagnetic interference and noise in the SQUID.

EXPERIMENTS WITH PLASTIC DISPLACER-REGENERATORS

Our experiments are intended only to demonstrate feasibility

of specific materials in a specific refrigerating mechanism

which may have a possibility of satisfying the stringent

requirements of a SQUID magnetometer.

The first experiment was to build a singlestage displacerregenerator (Fig. 2) of the type invented by Stirling in 1816,

but using modern materials. The displacer was a nylon rod 9.6 mm diameter by 300 mm long. It operated inside a spun-glass

epoxy cylinder (designated G-10 by the National Electronic

Manufacturers Association) with 2.4 mm wall thickness. The

radial clearance between displacer and cylinder was about .07

.07 mm, and the length of the cylinder was 313 mm, allowing

a 12.7 mm stroke for the displacer, which was connected to a thin brass push rod running through a sliding rubber seal at one end of the cylinder. The push rod was operated by a scotch yoke at the end of a shaft which was turned at 1 Hz by

a small electric motor. A pair of cams, 90° out of phase, on

the same shaft, operated the inlet and exhaust valves. (5) The

working fluid was helium gas, supplied at 14 atmospheres from

a storage cylinder and exhausted to the atmosphere when ex- panded. Twenty-five layers of aluminized plastic sheet were

wrapped around the displacer cylinder, which was mounted in an evacuated metal cylinder. Operated at 1 Hz, the machine

attained a temperature in the neighborhood of 50 K three or

four hours after starting, and maintained the temperature for

about two days on a cylinder of helium.

23
Helium gas from regulator /

Common shaft to cams and scotch yoke r / n t ^&ntake valve Cam

operated'

Exhaust valve

Solid nylon displacer

'6-10" Epoxy glass cylinder

Aluminized plastic radiation shields

Press fit and cement

Vacuum wall —

Pushrod to scotch yoke

Rubber seals

t Narrow radial gap regenerator

Figure 2 Details of simple cryocooler operating from com- pressed helium gas cylinder. Radial gap around

displacer is too narrow (^0.05 mm) to be shown in

this drawing. Displacer was 9.5 mm diameter by

300 mm long, with 12.7 mm stroke.

The success of this simple apparatus in producing rather low

temperatures encouraged us to construct a 3-stage, closedcycle, split Stirling machine, (6) using the same materials and

techniques (Fig. 3). Dimensions of the stages were as follows: The first stage displacer-regenerator was 245 mm long

and 19 mm diameter, with 0.05 mm radial clearance inside the

spun-glass epoxy cylinder. The second stage was 143 mm long

and 9.45 mm diameter with 0.07 mm radial clearance. The third

stage was 144 mm long and 4.7 mm diameter with 0.04 mm radial

clearance. The cylinders all had a 2.4 mm wall thickness and were epoxied together using a commonly available epoxy glue.

A brass cap was epoxied on the bottom or cold end. Cylindrical aluminum radiation shields were fitted over the lower end

of the first and second stages. Several layers of aluminized

plastic sheet were wrapped within the shields and around the

outside of the assembly for additional radiation shielding.

A polished stainless steel cylinder containing a 35.6 mm

diameter plastic piston formed the compressor. The various

dimensions of these machines were dictated by intuition and

by immediate availability of materials. In subsequent work

we hope to optimize the size and shape on a more rational

basis. The piston was driven through a 38 mm stroke by a

24
1 m Connecting line

Pushrod to scotch yoke

Rubber

piston ring

Felt grease ring Press fit and cement

Piston rod to crosshead

and crankshaft

Radiation shields

Vacuum wall —

Figure 3 Three-stage Stirling cryocooler. Displacer sections were 19 nun, 9.5 mm, and 4.7 mm diameter, and 245 mm,

143 mm, and 144 mm long, with 12.7 mm stroke. Piston

was 35.6 mm diameter with 38 mm stroke.

crankshaft and crosshead mechanism adapted from a small commercial freon compressor. Operating at 1 Hz with a displacer

stroke of 12.7 mm and a pressure excursion of about 3 to 13 Pa,

this machine maintained a temperature below 16 K. This machine was also operated

for the first cryocooler, and

temperature of 16 K. The perf

the dead space around the disp

displacer had a larger coeffic

glass epoxy cylinder, so that

"dead space" at the cold end a achieved. We therefore carefu

contraction and machined the d

the displacer and cylinder fit

state operation, rather than a

with the same valved cycle used

reached almost the same minimum

ormance was improved by reducing

lacer. We found that the nylon

ient of expansion than the spunthere were several millimeters of

fter steady-state operation was

lly measured the differential

isplacer accordingly, so that

ted together precisely in steadyt room temperature. After

25
initial cooldown, the dead space was eliminated by readjusting

the length of the push rod. When this was done, and the phase

angle between piston and displacer was empirically optimized,

the cryocooler maintained a temperature around 13 K or less,

for continuous periods as long as 5 weeks. The limit on nonstop operation appeared to be due to air being entrapped by

grease on the piston ring and the pushrod seal and contaminating the helium gas. Such contamination would eventually

build up, reduce the radial clearance of the displacer, and cause friction. We eliminated most of the contamination by

maintaining an ambient helium atmosphere on the back side of

the piston, but have not yet done so for the pushrod seal.

This three-stage displacer has been operated a total of nearly

5000 hours with no obvious signs of wear. It will maintain a temperature between 12-1/2 and 13 K at the cold end with no heat load. With a heat load of 11 mW the temperature rises to

14.5 K. The outer radiation shield is maintained at 120 K and

the inner radiation shield at 40 K. Operated at f = 1Hz with

a pressure excursion as noted above, the mechanical power re- quired for the compression stroke, assuming an isothermal

process, is PjV.f (lnP

2 /P1

) = (3 x 10 5

) (38 x 10" 6

) In (13/3)

watts = 15 W, where V. = 38 cm is the piston displacement.

Since a considerable fraction of the work is returned on the

expansion stroke, the net power required was estimated to be

of the order 10 W or less. This estimate was confirmed by

operating a similar machine with a 50 watt motor whose efficiency was less than 25%.

We have recently made fairly extensive measurements and

calculations of the various components of intrinsic heat load

in this machine, namely, shuttle heat, conduction, radiation,

and regenerator inefficiency, for the purpose of optimizing

the design. In this connection, we may wish to complicate the

design slightly by incorporating a high-heat-capacity matrix

at the cold end of the displacer cylinder, to improve the

regenerative heat exchange and achieve lower temperature. The

results of our measurements and the optimization will be

submitted for publication subsequently.

CONSIDERATION OF COMPATIBILITY WITH A SQUID MAGNETOMETER

Having demonstrated that superconducting temperatures (highTc

) can be achieved and maintained for long times with a very

low-power cryocooler made of plastics with favorable magnetic

properties, it remains to be seen whether such a machine can

be used with a highly sensitive SQUID magnetometer. Ignoring

for the present the problem of the compressor and its drive

mechanism, consider the magnetic signals generated by just

those parts that would be near the SQUID, namely the bottom

end of the displacer and its cylinder. In our experimental

26
machine, the end of the displacer is 4.7 mm diameter and moves

up and down about 13 mm. In the field of the earth, 0.5 x

10""4 t, this gives a peak-to-peak magnetic moment at 1 Hz of

about 0.3 cm

3 x 0.5 x 10~ 4 T x 1.1 x 10~ 5 = 0.16 x 10~ 9 T cm

3

. The susceptibility of nylon, 1.1 x 10"^, was measured for us

by F. Fickett.

If a SQUID were mounted 5 cm (for example) from the end of

the displacer, the peak-to-peak field at 1 Hz would be about

-12

10 ' ' T. A coherent oscillation of this magnitude could be

subtracted from the output with an accuracy of perhaps 10"^-^ T,

which is comparable to present SQUID sensitivities in a 1 Hz

bandwidth. Thus, one may be able to cope with the magnetic

signal from a small reciprocating displacer of nylon or com- parable material.

On the other hand, a peak-to-peak pressure excursion of 10 Pa

5 ^

(10 Pa = 1 atmosphere) in the displacer cylinder will cause

it to bend owing to non-uniform wall thickness and other

defects. If the SQUID were mounted on the end of the cylinder

for mechanical support, bending of the cylinder in the earth's

field may cause a flux change in the SQUID far too large to be

subtracted from the output with sufficient accuracy. There is no need to use precise numbers to show this. An elastic

modulus of 10 Pa and a pressure change of 10 Pa in a cylinder whose ratio of length to diameter is 10 gives a bend of

the order of 10 x 10 6 x 10

-11

x 10~ 2 = 10~ 6 radians, if the

lateral inhomogeniety (different wall thickness, for example)

is of order 10~2. in the earth field this would appear as an oscillation of the order of 10~10 t. Two alternative approaches

suggest themselves for reducing this signal. One is to measure the bending and empirically correct it by grinding or scraping one side of the cylinder. The other is to mount the

SQUID on an independent support with only a flexible thermal

connection to the cold end of the cylinder. In any case, it

may be quite difficult to reduce or compensate the bending

signal to a level of 10~ 14 T. COMPARISON OF THE CRYOCOOLER WITH CONVENTIONAL LIQUID-HELIUM

CRYOSTATS

We have shown that the lowpower cryocooler can maintain temperatures below 15 K with ten or twenty milliwatts heat load.

Bias power requirements for SQUIDs are many orders of magnitude smaller than this, and the only significant heat load

imposed by the SQUID is conduction down the electrical leads.

Our experience with conventional liquid-helium cryostats

indicates that, with good design, the conduction heat load can

be reduced to less than ten milliwatts. Actually, designing

or specifying cryostats on the basis of refrigeration capacity

at the coldest point is quite misleading. What is important

27
is to distribute the refrigeration capacity, using a multistage machine, so that the bulk of the heat input (conduction,

radiation, etc.) is pumped at relatively high temperatures,

leaving very little to be pumped at the lowest temperature.

The distribution of refrigeration capacity vs temperature is

another aspect of design optimization which has received

little attention in the literature. For comparison, a typical

small liquid-helium cryostat with an evaporation rate of one

liquid liter per day (corresponding to about 700 liters of gas

at room temperature) is supporting a total heat load, intrinsic plus imposed, of 29 mW at 4.2 K, and has the capability of

absorbing an additional 7.6 m W/K due to the heat capacity of

the vapor, which is approxmately independent of temperature.

These figures might be used as a guide for the necessary

refrigeration capacity of a small cryocooler. Actually, the

figures should be quite conservative because (1) a cryocooler

would not have the large cold reservoir and (2) optimum design

would probably not require the temperatureindependent heat

absorption capacity, but rather would concentrate the refrigeration capacity at higher temperatures.

FURTHER PERTINENT COMMENTS

The work we have reported here is unique in two respects,

first in the choice of materials and construction methods for

the displacer, and second in building a machine with at least

an order of magnitude lower input power requirements than

previous comparable machines. Both of these features are, of

course, related to our particular aim of minimizing magnetic

interference, but they will have considerable effect on cost

and portability as well. It appears from our experiments that

the cost of a cryocooler could easily become comparable to the

cost of a SQUID magnetometer or similar superconducting instrument. Such systems might well be justified economically,

in view of the fact that the cost of liquid helium will in- crease as the supply diminishes, whereas the cost of cryocoolers should decrease substantially if they are produced in

sufficient numbers to replace present liquid-helium cryostats.

This work completes what might be called a preliminary study

of the feasibility of using a small mechanical cryocooler for

SQUID magnetometers and other small superconducting devices.

We conclude that a second phase of development is justified.

We believe this should include (1) optimizing the design as noted above, (2) incorporating a high-Tc SQUID with which to measure and evaluate the level of magnetic interference from

the various moving parts, and (3) evaluating other potentially

useful materials and refrigeration processes which may improve

the performance.

28
ACKNOWLEDGEMENT

We are grateful to several individuals for helpful comments

and criticism, particularly Sam Collins, Dave Daney, T. R. Strobridge, and R. 0. Voth. This work is supported jointly by

the National Bureau of Standards, the Office of Naval Resarch

and the Naval Electronics Systems Command.

REFERENCES

1. For example, we have built 1.7-liter spun-glass-epoxy

cryostats with hold times of nearly three days on one filling of liquid helium. At the opposite extreme, a 180-liter cryostat has been built with a hold time of

102 days. See R. J. Dinger, J. R. Davis and M. Nisenoff,

NRL Memorandum Report 3256, Naval Research Laboratory,

Washington, DC (March, 1976).

2. Some contrasting comments on reliability by different

people (not directed specifically at each other) may

be found in Proc. Cryogenic Cooler Conf., AFFDL-TR73-149, Air Force Flight Dynamics Laboratory,

Wright-Patterson Air Force Base, Ohio 45433 (Dec.

1973). This report also contains several papers on recent developments in miniature cryocoolers for

the 10-20 K range.

3. R. C. Longsworth, p. 390, Proc. Fifth International Cryogenic Engineering Conf. , IPC Science and Technology

Press (1974), has shown that liquid-helium hold times

can be increased by a factor of 4 or more by maintaining a shield at 20 K rather than at 77 K. He estimates that one might achieve hold times of the order of

a year with a 25-liter cryostat. His results should be

quite conservative as applied to cryostats for lightweight low-energy devices like SQUIDs.

4. Much historical and recent work on Stirling and

related machines is described by G. Walker, StirlingCycle Machines. Oxford University Press (1973).

5. Operation of a Stirling-type displacer-regenerator

with a valved helium supply was developed by H. 0.

McMahon and W. E. Gifford, p. 354, Advances in Cryogenic Engineering, Vol. 5, Plenum Press (1960).

6. The stepped-displacer idea used here was invented

by Fokker and Kohler. The equivalance of a steppeddisplacer machine to a multi-stage machine is ex- plained in some detail by G. Prast, Philips Tech.

Rev. 26, 1 (1965). The term "split" refers to the

fact that the compressor cylinder and displacer

cylinder are geometrically separated from each other.

29
Conference of the German Refrigeration Society, Munich (October 1976).

REFRIGERATION FOR SMALL SUPERCONDUCTING DEVICES*

J. E. Zimmerman, R. Radebaugh and J. D. Siegwarth

Cryogenics Division

National Bureau of Standards

Boulder, Colorado 80302

ABSTRACT

The present state of the art is such that the greatest cost in

using small superconducting devices, in many applications, is the ex- pense and inconvenience of the associated cryogenic system. The prac- ticality of such devices would be greatly enhanced, therefore, if a practical and economical self-contained, closed-cycle cryocooler could

be made compatible with such devices. We have built an experimental

cryocooler whose unique features (non-magnetic materials and very low

input power) may represent a partial solution to the problem.

Small superconducting devices are currently operated almost universally in liquid-helium cryostats. Evaporation rates vary from less than

a liter per day to several liters per day, depending upon the size and

design of the cryostat and the heat load imposed upon it. Continuous

operating times as long as 102 days have been achieved with a 180- liter

cryostat built for a SQUID magnetometer , which may be taken as the pro- totype for practical superconducting devices. Because of the increasing

number of applications of such devices, we have begun a study of the

possibility of maintaining the required operating temperatures with a low-power closed-cycle cryogenic refrigerator, which may have advantages

of portability, versatility, and long continuous operating time. Since

the cost of helium will increase as the supply diminishes, and the cost

of closed-cycle refrigerators should decrease substantially if they are produced in sufficient numbers to replace present liquid-helium cryostats, the shift of technology can be economically justified.

In the first part of this paper we emphasize that a need exists for

a low-power, closed-cycle cryogenic refrigerator (cryocooler) for the

Contribution of the National Bureau of Standards,

not subject to copywrite.

30
convenient operation of a number of small superconducting devices, which

are now or may soon be used in a broad spectrum of applications, and in

the second part we describe briefly our own preliminary experiments

directed toward satisfying these needs.

Several superconducting devices with uniquely useful properties and

an impressive range of potential applications have appeared in the last

few years. A bolometer, operating at the superconducting transition

temperature, is an order of magnitude more sensitive (noise-equivalent

power 2 x 10 W/^Hz ) than any other device of its type . Similarly, a so-called super-Schottky diode, a superconducting Pb contact on GaAs,

exhibits an X-band noise temperature less than 10K, a hundred times

better than conventional room-temperature diodes . The list of applications of the Josephson effect is too long to be summarized here, and we will only mention two. A microwave-biased Josephson junction has re- placed the electrochemical cell as the primary voltage standard in the

leading standards laboratories of the world. Superconducting quantum

interference devices (SQUID' s) in many sizes and shapes are used in hundreds of laboratories as magnetometers, galvanometers, susceptibility

meters, dilatometers, rf power meters and rf attenuation standards, vlf

4 communication receivers, and many other purposes . In terms of sheer number of uses and users, the SQUID magnetometer

is clearly the instrument with greatest practical impact at the present

time, compared to other superconducting devices. In addition to numer- ous applications in low- temperature physics laboratories, it is being

used by significant numbers of specialists in entirely different fields,

c c 7 e.g., geology , medical research , and military purposes . In all such

applications, the users accept the inconvenience and cost of a cryogenic

system because a SQUID magnetometer is a simple and rather inexpensive

-14 i— instrument, and it is more sensitive, e.g., 10 T//Hz, than any comparable conventional instrument by several orders of magnitude.

All or most of the devices listed above are so superior, in sensitivity and other important respects, to conventional room-temperature

devices, that there is little doubt they would predominate in practically all applications, were it not for the associated cryogenic system.

31
Both the first cost and the operating cost of present cryogenic systems

usually far exceed the cost of a device like a SQUID or a superconducting bolometer. Thus, it is the cryogenic system which currently limits

the acceptance of these devices, not the characteristics of the devices

themselves. Viewed in this light, it is seen that an order of magnitude

improvement in the cryogenic system (involving such qualities as cost,

operating convenience, size, reliability, etc.) might well result in an enormous enhancement of the user acceptance.

Another feature common to the devices listed above is that their

intrinsic refrigeration requirements are microscopic. The bias power, essentially all of which is dissipated and must therefore be absorbed

—10 —8 by the cryostat, varies from 10 W for the bolometer, to 10 W for the

-12 diode, to 10 W for an rf-biased SQUID. Room- temperature radiation to surfaces with an absorptivity of say 2% would amount to about 10 W,

2 assuming that the cold device has a surface area of the order of 1 cm . It is difficult to make any general estimates of heat conduction along

the mechanical supports and electrical leads, since design requirements

might be quite varied. An empirical upper limit of refrigeration required is obtained from our experience with the operation of SQUID magnetometers in liquidhelium cryostats. We have observed a difference of evaporation rate of

the order of 0.1 t/day with and without the SQUID in the cryostat,

corresponding to an extra heat input into the liquid helium of 3.5 mW

due to the SQUID. It is not our purpose here to develop careful theoretical estimates of the intrinsic refrigeration requirements for small

superconducting devices. However, very rough estimates show in principle (not necessarily implying practicability) that an optimum system

should require several orders of magnitude lower refrigeration power

than any existing practical cryocooler is capable of providing. On the basis of all the experimental and analytical evidence, we felt it

would be useful to experiment with the design of a very- low-power closedcycle cryocooler, simply disregarding the small heat load imposed by the

superconducting device. Our multiple uses of a SQUID magnetometer led us also to consider the possibility that the cryocooler could have a sufficiently low level of magnetic interference as not to seriously degrade

32
the sensitivity of this particular instrument. Ideally, this means that

-14 the interference level should not exceed about 10 T, although in

specific applications a greater level might be tolerable. For niobium

point-contact SQUID' s in use at the present time the required operating

temperature is about 9K or below. Thin-film SQUID' s of high-T alloys

(8) C have been demonstrated to operate as high as 16K

We have built a laboratory model cryocooler, with the short-term

objective of mounting in it a magnetometer SQUID and experimentally

measuring the level of magnetic interference. The ultimate aim is to produce a practical self-contained cryocooler-magnetometer system. The

present model is a split Stirling machine (see Figure 1) , with piston

and displacer in separate cylinders connected by 1 m of 2 mm inside

diameter tubing. Swept volume of the piston is 38 cm , swept volume of

the displacer is 3.6 cm , and the displacer was made of solid nylon rods

with three cylindrical sections 19 mm, 9.5 mm, and 4.7 mm in diameter

(see Figures 2 and 3). Thus, during the expansion stroke of the piston,

refrigeration is produced at the two steps as well as at the small end,

the expansion volumes being 2.7 cm , 0.68 cm and 0.23 cm respectively,

3 (9) totaling 3.6 cm . The displacer cylinder was assembled from three

pieces of commercially available glass-reinforced epoxy tubing. The

displacer was carefully machined to fit the cylinder with less than

0.1 mm radial gap on all three sections. Regenerative heat exchange

takes place between the working fluid (helium) and the solid surfaces

along this gap, exactly as in Stirling's original engine of 1816. Thin tubular aluminum radiation shields were pressed onto the two steps on the cylinder. The piston and the displacer were coupled by

appropriate mechanisms to a common crankshaft, arranged so that the

phase relation between the two could be empirically adjusted for optimum

results.

This machine has been operated a total of about 5000 hours at 1 stroke per second with the helium pressure oscillating between 3 and

5 2 12 x 10 N/m . After some minor initial alterations it maintained a temperature at the small end of the displacer cylinder of 13K, well

within the operating requirement of a high-T SQUID. It required 50 W

input to an electric drive motor whose measured efficiency was less than

33
25%. Thus, the actual mechanical power input was approximately 10 W. This agrees reasonably well with the power calculated from the piston

displacement and the pressure excursion

We wish to emphasize that our cryocooler design represents only a tentative first step toward a possible practical machine for small

superconducting devices, the realization of which seems to us potentially very rewarding. We have demonstrated (1) that non-magnetic,

insulating materials such as nylon and glass-reinforced epoxy can be

used in reciprocating machines with rather long operating time, and (2) that useful cryogenic temperatures can be maintained with very low input

(12) power , if the superconducting system is compatible with very low refrigeration capacity.

This work is supported in part by the Office of Naval Research and

by the Naval Electronics Systems Laboratory.

34
REFERENCES

1. J. R. Davis and M. Nisenoff , International Conference on Superconducting Quantum Devices, Berlin (West), October 5-8, 1976. (Proceedings to be published by Walter de Gruyter, Berlin)

.

2. J. Clarke, Proceedings of the Berlin conference cited in ref. 1. See also Proc. 14th Int. Conf. on Low Temperature Physics 4^ 226

(North-Holland Publ. Co., Amsterdam, 1975).

3. M. McColl, et al, Appl. Superconductivity Conference in Palo Alto

(California) , August 1976 (To be published in IEEE Trans on magnetics)

.

4. Several of these applications were described at the Berlin con- ference cited in Reference 1.

5. N. V. Frederick, et al, IEEE Trans, on Geosci. Electronics GE-12,

102 (1974). J. E. Zimmerman and W. H. Campbell, Geophysics 40,

269 (1975).

6. D. Cohen, IEEE Trans on Magnetics MAG- 11 694 (1975)

.

7. W. M. Wynn, et al, IEEE Trans, on Magnetics MAG-11 , 701 (1975) (See

also Reference 1)

.

8. T. Fujita, et al, IEEE Trans, on Magnetics MAG-11 , 739 (1975).

9. The principle of Stirling machines with multiple expansion spaces

is described in detail by G. Prast, Philips Tech. Rev. 26, 1 (1975).

10. See for example G. Walker, Stirling-Cycle Machines, Oxford University Press (1973)

.

11. More details on the operation of this cryocooler and the possibility of operating a SQUID magnetometer with it were given at the

Berlin conference cited in Reference 1. 12. The lowest input power requirement for any previous practical

cryocooler has been given as about 300 W. See a survey by T. R. Strobridge, NBS Tech. Note 655 (June 1974)

.

35
•REGENERATOR

GAP

PISTON

Qi Rejected

Absorbed

REGENERATOR

HEAT CAPACITY

(Q 3 = Qb for Perfect Gas)

Figure 1. The essential elements of a Stirling machine, and the theoretical thermodynamic cycle (from ref . 11)

1 m Connecting line ^^1S_

{ Pushrod to scotch yoke

Rubber

piston ring

Felt grease ring Press fit and cement

Piston rod to crosshead and crankshaft

Radiation shields

Vacuum wall —

Figure 2. Sketch of an experimental three-stage Stirling cryocooler

(from ref. 11)

36
«tt

Figure 3. Photograph of disassembled cryocooler. Left to right:

working cylinder, piston, displacer cylinder, two radiation

shields, and displacer. Fittings were provided for maintaining a helium atmosphere on the non-working side of the

piston (tube at center left) , connecting a pressure transducer

(top left) , and filling the system with helium, (top center.) .

37
"Applications of Closed-Cycle Cryocoolers to Small Superconducting Devices," NBS Special Publication

SP-508 (J. E. Zimmerman and T. M. Flynn, eds.)> Superintendent of Documents, U. S. Government

Printing Office, Washington, D. C. 20234 (1978), pps. 59-65.

Operation of a SQUID in a Very Low-Power Cryocooler *

J.E. Zimmerman and Ray Radebaugh

Cryogenics Division

National Bureau of Standards

Boulder, Colorado 80303

Abstract

A point-contact Nb SQUID has been operated a total of several hundred hours

at a temperature of about 8.5 K in a low-power Stirling cryocooler with a fourstage displacer. The system requires the order of 15 W of mechanical drive

power. Except for the drive motor, the entire unit is non-magnetic, and the

displacer and cold cylinder are made of non-conducting plastics, to minimize

ferromagnetic and eddy-current fields, which would interfere with the operation

of the SQUID when used as a magnetometer. With the system operating in the

-5 -9 earth's field B , an ac interfering signal of about 10 B (i.e., -v 10 T) at the SQUID was seen at the 1 Hz operating frequency of the cryocooler. This

signal was probably caused by rotation of the SQUID in the earth's field due to pressure flexing of the cold cylinder, although magnetic impurities in the moving

parts would also contribute. Measurements of the complete self-generated interference spectrum, and experiments aimed at reducing interference to levels of

-12

10 T and below are in progress.

We have previously published an essentially theoretical discussion of the operation of a very- low- power superconducting device, such as a SQUID, in a closed-cycle cryogenic refrigerator (cryocooler) [1,2]. It was pointed out that the refrigeration capacity required for

the devices themselves is negligible, and the capacity required to cool the electrical leads

is small. On the other hand, a SQUID magnetometer, in particular, places a rigorous upper

limit on the allowable level of magnetic and mechanical interference.

It is assumed, of course, that all of the major heat loads (room temperature radiation,

and heat flow along electrical leads and support members, for example) are intercepted and

pumped out at one or more higher-temperature refrigeration points, rather than at the cold

end. In fact, in most small-device applications, it would seem that the refrigeration re- quirement at the cold end is almost incidental, if the heat loads at the higher-temperature

points are handled efficiently. Obvious though this point is, cryocoolers are very often

described in terms of cold-end refrigeration capacity only.

In liquid-helium cryostats, higher-temperature refrigeration is provided by the helium

vapor, whose total refrigeration capacity, uniformly distributed in temperature, is about 70

times that of the evaporating liquid itself. In a cryocooler one has the opportunity, in principle, of optimizing the distribution of refrigeration at several points between ambient

and low temperature regions. Furthermore, a cryocooler is required to cool only the device

* Contribution of the National Bureau of Standards, not subject to copyright.

Supported in part by the Naval Electronics System Command.

38
itself and not a large liquid reservoir. Thus, for two reasons, estimates of refrigeration

requirements based on experience with liquid-helium cryostats may be quite conservative.

Since there is a pervasive casual tendency to specify cryocoolers only by their refrigeration capacity at cold end, it may be useful to illustrate this point with a hypothetical,

but fairly typical, example. A one-liter plastic liquid-helium cryostat with a single vaporcooled radiation shield at a steady-state temperature of 150 K might have a heat leak to the

helium, as measured by the evaporation rate, of 30 mW. Here the evaporation rate and the

shield temperature are mutually interdependent. That is, a lower shield temperature would

reduce the evaporation rate, but then the flow of cold vapor would not maintain the shield

temperature at the lower value. On the other hand, if a cryocooler were used to cool the same one-liter volume and the same radiation shield, it could be designed to maintain the

shield at 50 K instead of 150 K. If the heat leak to the one-liter volume were primarily

thermal radiation, then it would be reduced by a factor of 81, from 30 mW to about 0.4 mW.

If there were appreciable thermal conduction as well as radiation, the comparison would not

be quite so dramatic, but still impressive.

We also described [1,2] an experimental cryocooler particularly suited for operating

very low-power high-T superconducting devices. This cryocooler was a three-stage split

Stirling-cycle [3] machine, which was operated more than 6000 hours at 1 stroke-per-second,

and reliably maintained a temperature of 13 K or less for uninterrupted periods of up to 5 weeks. Its novel features were (1) non-magnetic, non-conducting plastics for the displacer

and cold cylinder, (2) low operating power, and (3) mechanical simplicity. Solid nylon rod was used for the displacer and commercial spun-glass epoxy ("G-10") tubing for the cold

cylinder. The input power of 50 W to an electric drive moter with about 15 W mechanical

output power was a factor of 10 or more smaller than the power requirements of any previous

machine for this temperature range, so far as we are aware. Regenerative heat exchange

occurred in the narrow radial gap between the nylon displacer surface and the inner surface

of the cold cylinder. Gap regeneration was used in Stirling's original engine in 1816, and,

being mechanically trivial, is commonly used in miniature engines described in hobbyists

magazines.

A machine similar to ours (plastic displacer, gap regeneration, and very low power) was

built several years ago by du Pre and Daniels [4]. They achieved 130 K with a two-stage

displacer and 98 K with a conical displacer. The latter may be regarded as the limiting case

of a displacer with an infinite number of stages.

The temperature maintained by our three-stage machine is well within the range of high-T

superconductors such as NbN and Nb

3 Sn. Practical devices have been made to operate at 14 K

and above [5].

A new four-stage displacer and cold cylinder (Figures 1, 2 and 3) were built to replace

the three-stage unit. With this modification and some minor refinements, the machine has

maintained around 8.5 K for a total of several hundred hours of operation. The four sections

of the nylon displacer were, respectively, 4.7, 9.5, 19 and 28 mm diameter and 15, 10, 12 and

12 cm long. The cold cylinder was made from sections of commercial spun-glass epoxy tubing

("G-10") with 2.4 to 4 mm wall thicknesses. The displacer sections were carefully

machined to fit the cylinder with almost no radial clearance, except for a few cm of the

39
large end where about 0.1 mm radial clearance was provided. During cooldown, differential

contraction of the nylon rods and the glass-epoxy tubes provides about one percent radial

clearance, except at the large (warm) end. This technique of fitting the displacer to the

cylinder is quite simple, and it also gives the minimum possible regenerator gap volume, which

is important for achieving minimum temperature. As in the earler model, the exact lengths of

the sections were calculated to take account of differential contraction, so as to give precise longitudinal fit in each section after cooldown. Here the fitting technique is not so simple. The important thing is to avoid any appreciable dead volume at the small (cold) end.

Slight misfitting at the larger steps seemed to have little effect on the cold-end temperature.

Our measurements of the various components of stray heat inputs, including "shuttle

heat," reported separately [6], led us to conclude that lower temperature would be achieved by

reducing the displacer stroke. Thus, the four-stage displacer has been operated with a 7 mm

stroke, as compared to the 12 mm stroke of the three-stage model. Other operating parameters were essentially the same for both models, namely: operating speed 1 Hz, average pressure

3 0.5 MPa (5 atm), and piston displacement 38 cm . The average pressure, the speed, and the

displacer stroke could be varied considerably around these values without changing the coldend temperature very much. It was found that the values of these three parameters giving

lowest ultimate temperature were generally lower than the values giving fastest cooldown --

a result not entirely unexpected.

The machine was mostly operated with the cold end down. However, it was also operated

for a few hours with the cold cylinder horizontal, and also with the cold end up. In the

latter case, the steady-state temperature was about 0.2 K greater than for the other two positions, e.g., 8.7 K instead of 8.5 K. We did not investigate the reason for this small

gravitational effect.

Temperature of the first (large) step was measured with a thermocouple, and the cold-end

temperature was measured by a helium gas bulb clamped directly to the small (1 cm diameter)

end of the cold cylinder. A simple permanently-adjusted point-contact rf-biased Nb SQUID was clamped to the gas bulb. Low-conductivity parallel wires were used as rf coil leads, rather

than the usual copper coaxial line, in order to reduce heat leak. Temperature of the first

step stabilized at values of 180 K for average pressure 0.5 MPa to 150 K for average pressure

0.7 MPa. The lowest stable cold-end temperature was obtained at about 0.5 MPa, as noted

above.

Our present data on magnetic and mechanical interference are meager. With the SQUID

oriented North/South in the earth's magnetic field, there was a 0.5 nT ac component in the

output at the 1 Hz operating frequency of the cryocooler. This was most likely due to pres- sure flexing of the cold cylinder, which would result from lateral inhomogeneities in wall

thickness or elastic properties. Other lower-amplitude noise components at higher frequencies

4 have not yet been measured or identified. The 0.5 nT interference at 1 Hz is only about 10

times greater than typical SQUID sensitivities. If the hypothesis of pressure flexing is correct, then this component could perhaps be reduced by a factor of 100 by correcting the

Inhomogenieties in the cylinder walls. Another factor of 100 could probably be achieved by

accurately measuring the wave shape and amplitude and subtracting it from the output. In any

40
case, this particular Interference would not so seriously affect measurements at frequencies

other than 1 Hz. Measurements of the full self-generated Interference spectrum for the system remain to be

done. At this point we believe that a useful low-cost system can be built for magnetic measurements in the picotesla range. Similar systems could be built for other device applications [2]. The Josephson voltage standard [7] seems a particularly good candidate.

Just as the point contact made Josephson devices such as SQUID'S available to everyone

who took the time to understand a simple but unfamiliar principle, the work reported here

demonstrates that superconducting temperatures are also available to everyone who takes the

time to acquire modest skill with a lathe.

References

[1] J.E. Zimmerman, R. Radebaugh and J.D. Siegwarth, "Possible Cryocoolers for SQUID Magnetometers," Superconducting Quantum Interference Devices and Their Applications . H.D. Hahlbohm and H. Lubbig, eds., Walter de Gruyter, Berlin (1977) p 287.

[2] J.E. Zimmerman, R. Radebaugh and J.D. Siegwarth, "Refrigeration for Small Superconducting

Devices," Conf. German Refrigeration Society, Munich (October 1976) (in press).

[3] G. Walker, Stirling-Cycle Machines , Oxford University Press (1973).

[4] F.K. du Pre\ A. Daniels, "Gap-Regeneration Method for Stirling and Similar Cycles,"

Progress 1n Refrigeration Science and Technology, Vol. 1, Proc. XIII International

Congress of Refrigeration, Wash. D.C., (1971) AVI Publishing Co., Inc. Westport, Connecticut, p 137 (1973).

[5] M. Beasley in this publication.

[6] R. Radebaugh and J.E. Zimmerman, in this publication.

[7] L. Holdeman, in this publication.

41
Plastic Piston

in Cylinder

Helium

Gas

Glass Reinforced

Plastic Cylinder

First Step

Vacuum Space

Aluminum

Vacuum Can

Radiation Shields

Cold End

Squid

Figure 1. Schematic diagram of low-power SQUID cryocooler system, not to scale. Displacer dimensions and other data are given in text.

42
Figure 2. Cold cylinder and displacer of 8.5 K cryocooler. Left to right: 30 cm scale,

nylon displacer, glass-reinforced epoxy cylinder with inner radiation shields in place, outer radiation shield, and vacuum jacket.

43
Figure 3. Room-temperature end of cold cylinder assembly. Clockwise from left: gas line

to working cylinder, gas-thermometer line, thermocouple wires, vacuum connection,

and rf SQUID connector.

44
"Applications of Closed-Cycle Cryocoolers to Small Superconducting Devices," NBS Special Publication

SP-508 (J. E. Zimmerman and T. M. Flynn, eds.), Superintendent of Documents, U. S. Government

Printing Office, Washington, D. C. 20234 (1978), pps. 67-73.

SHUTTLE HEAT TRANSFER IN PLASTIC DISPLACERS AT LOW SPEEDS *

Ray Radebaugh and J. E. Zimmerman

Cryogenics Division

National Bureau of Standards

Boulder, Colorado 80303

ABSTRACT

Previous analyses of shuttle heat transfer in refrigerators with displacers have neglected

radial temperature gradients in the displacer or cylinder. Such analyses are valid when the

gas gap 1s the dominant thermal resistance. We show that with plastic materials for the

displacer and cylinder, shuttle heat transfer can be dominated by the thermal penetration

depth of the plastic when operating at low speeds. An equation is derived for the shuttle

heat transfer 1n such cases. Experimental data on shuttle heat transfer is obtained for

temperatures down to 120 K for different strokes and speeds and the agreement with calculated

values 1s good.

* Supported in part by the Naval Electronic Systems Command

45
SHUTTLE HEAT TRANSFER IN PLASTIC DISPLACERS AT LOW SPEEDS

Ray Radebaugh and J. E. Zimmerman

Cryogenics Division

National Bureau of Standards

Boulder, Colorado 80303

INTRODUCTION

The Stirling-cycle refrigerator described previously [1] requires no net refrigeration

power to keep a SQUID cold. The lowest temperature reached 1s the point where the total

refrigeration power 1s cancelled by various heat flows to the low temperature end. These

heat flows come from several sources, Including radiation, heat conduction down the dlsplacer,

regenerator Inefficiency, and shuttle heat transfer. Radiation heat flow to the lowest

temperature can be made negligible in most cases with good shielding. For temperatures

somewhat above 10 K the regenerator Inefficiency can be made fairly small. Thus the major

heat flow 1n much of the refrigerator Is from the conduction and shuttle processes. Shuttle

heat transfer occurs when the dlsplacer, with an axial temperature gradient, reciprocates

Inside a cylinder with a similar temperature gradient. Because shuttle heat transfer 1s a major source of heat leak, it must be well characterized in order to optimize the design of

the refrigerator.

PREVIOUS WORK

Zimmerman and Longsworth [2] discussed shuttle heat transfer and derived an equation for

1t based on a simplified theory. Their theory contained four simplifying assumptions: (1)

properties are only weak functions of temperature, (11) heat capacity of the dlsplacer and

wall are Infinite, (111) negligible gas pressure cycling, and (1v) Interaction of axial

conduction effects 1n the dlsplacer are neglected. For square-wave motion of the displacer

they derive the equation

V Tc (i)

for the shuttle heat transfer, where k 1s the thermal conductivity of the gas, D is the

dlsplacer diameter, S 1s the stroke length, t 1s the radial clearance, L 1s the length of the

dlsplacer, T

h 1s the hot temperature, and 1

Q

1s the cold temperature. For sinusoidal motion

the average temperature drop across the gas gap 1s 2/tt times that for square-wave motion.

Hence for sinusoidal motion the average shuttle heat transfer becomes

<."¥ vV

(2)

More exact calculations may give different numerical doefficlents, but the functional dependence of the shuttle heat loss on the various parameters will remain the same. Harness and Neumann [3] calculated shuttle heat transfer when assumption (1v) was relaxed. The problem then becomes very complex and can be solved only by numerical Integration

of a second-order differential equation. Their results for the total heat flow differ from

46
that of adding the static heat conduction down the displacer to the simple shuttle heat

transfer equation given above. The difference ranges from 5 to 40%, depending on various

parameters, and is not always of the same sign. As far as we know, attempts to solve the shuttle heat transfer problem with any of the

other three assumptions relaxed has never been done. However, we were concerned that assumption (ii) was not satisfied when the displacer and wall materials were of plastic and their

operation was at the low speed of about 60 rpm. When the heat capacity of the displacer and

wall are assumed to be infinite, radial temperature gradients are thus neglected. In highspeed operation little heat is transferred across the gas gap in each cycle. As a result,

the surface temperature of the displacer and wall remain nearly constant in time and the

assumption (ii) is satisfied. Temperature cycling of the surface does become important at

low speeds. The use of plastic materials instead of metal increases the amount of temperature cycling because the lower thermal conductivity decreases the thermal penetration depth.

THEORY

In order to keep the problem from becoming too complex, we first make the following

assumptions: (i) properties are nearly constant with temperature, (ii) thermal resistance of

the gas gap is negligible, (iii) the effect of gas-pressure cycling is neglected, i.e.,

regenerative action is perfect and heat capacity of the gas in the gap is negligible, (iv)

interaction of the axial conduction effects in the displacer and wall are neglected, and (v)

motion of the displacer is sinusoidal.

The shuttle heat transfer is just the net enthalpy flow resulting from the motion of the

displacer. Its temperature is higher during the hot-to-cold motion than during the reverse motion so enthalpy is transported toward the cold end. The net enthalpy flow is just equal

to the heat transferred to the displacer during one-half the cycle over a length equal to the

stroke. To evaluate this heat transfer we first let the wall have infinite heat capacity so that it has no temperature cycling. We relax this condition later. Because of assumptions

(ii) and (iii) above, we can say that the surface of the displacer takes on the temperature

of the wall adjacent to it at any moment in time. The sinusoidal motion of the displacer

then causes a sinusoidal temperature fluctuation of the surface, provided the axial temperature gradient is constant over the distance of the stroke. The problem of sinusoidal tem- perature fluctuations on the surface of a semi-infinite solid is commonly treated in texts on heat conduction. Schneider [4] treats such a problem and the total heat transferred during

one-half of the cycle is given as

Q = 2AAT /kTC , (3)

where A is the surface area, AT is the peak of the temperature fluctuation, k is the thermal

conductivity of the displacer, C is its specific heat per unit volume, and w is the frequency

of oscillation. For n cycles per second o» = 2TM. Now A = ttDS and AT = (dT/dx)(S/2). Thus

the heat transfer during one-half cycle and the net enthalpy flow per cycle become

47
Q = DS 2 SiU/I5t ( 4 )

If the wall material has the same specific heat and thermal conductivity as that of the

displacer, then AT will be just one-half of t\

transfer with n cycles per second then becomes

displacer, then AT will be just one-half of that considered above. The average shuttle heat

ft - timlnvW /»rkC _ DS 2 (dfl /rmkC

Q

s - (l/2)nDS [dYjW-^7 " — [d^J V —mrkC

(5)

As before, the functional dependence should be correct, but more exact calculations may give

slightly different numerical coefficients.

For a thermal conductivity independent of T we can write dT/dx = T, - T_ /L. The

primary difference between this shuttle heat transfer and that derived by Zimmerman and

Longsworth is the vn dependence in the present calculation. Thus at low speeds the thermal

resistance of the solid material limits the shuttle heat transfer, whereas at higher speeds

the thermal resistance of the gas gap limits the shuttle heat transfer and Eq. (2) must be

used. For intermediate speeds we derive the total shuttle heat transfer by considering the

two thermal resistances to be in series. Hence

i/0

s « i/09

s + i/0

s

s

, (6)

where tip and ?) s are the shuttle heat transfers limited by the gas gap and by the solid

material, respectively.

Equation (5) will be valid as long as the radius of the displacer and the wall thickness

is somewhat greater than the thermal penetration depth in the material. The thermal penetration depth X is defined as that depth where the amplitude of temperature fluctuation is 1/e

of that at the surface and is given approximately by [4]

X = \/k/rmC , (7)

where k/C is just the thermal diffusivity. Table 1 shows values of k, C, and X for nylon at various temperatures with n = 1 Hz. Values for k are from Ashworth, et al . [5], and values

for C are estimated from two different sources [6,7]. The specific heat and thermal con- ductivity of G-10 glass-reinforced epoxy is roughly the same as that of nylon [8]. From the

penetration depths given in Table 1 we see that finite size effects can be neglected in most

practical cases.

48
Table 1. Thermal conductivity, specific heat, and

thermal penetration depth of nylon for n = 1 Hz

T

(k)

k (mW/cm K)

C

3 (mJ/cnr K)

X(mm)

300 3.66 2000 0.24

200 3.40 1430 0.28

100 2.97 833 0.34

50 2.19 440 0.40

20 0.948 112 0.52

10 0.385 22 0.75

5 0.156 3.2 1.25

EXPERIMENTAL

To our knowledge previous measurements of shuttle heat transfer have not been made. It was felt that such measurements were necessary to check the accuracy of the calculations from

a simple theory. Shuttle heat transfer can be measured by first observing the warm-up rate of a known thermal mass at the cold end of the displacer cylinder. During this measurement

the displacer is operating but the compressor is not. Next, the displacer is shut off and

the static warm-up rate is measured. The difference in the two measurements gives the

shuttle heat transfer. Effects of gas pressure cycling are determined by varying the average

pressure 1n the system. In this paper measurements down to 120 K are reported and for this

temperature range the gas pressure had no effect.

Measurements were made using a Stirling-cycle refrigerator with a single stage displacer.

The displacer was made of nylon and the cylinder wall was of G-10 epoxy, 2.3 mm thick The

displacer was 0.95 cm in diameter and 30.3 cm long. The speed and stroke could be varied.

The radial clearance between the displacer and wall was about 0.1 mm at 145 K and 0.06 mm at room temperature. Some measurements were made with nitrogen gas instead of helium gas in the

system to increase the thermal resistance of the gas gap.

RESULTS

Figure 1 shows the shuttle heat transfer as a function of the cold-end temperature when

the warm end was at 300 K. The calculated values consider both the gas gap and penetration

depth limit using Eq. (6). Agreement between experimental and calculated values is excellent

for a stroke of 1.19 cm. For a 2.45 cm stroke the agreement is excellent at higher temperatures but worsens at lower temperatures. The disagreement may be a result of the temperature

gradient not being constant over the stroke length or it could be a result of interaction

with static thermal conduction.

49
08

06

S

.£ 0.4

02

1

1 SHUTTLE HEAT TRANSFER

He Gas Speed=54 rpm

Stroke-2 45 cm

• Experimental —Calculated

T, K

Figure 1. Measured and calculated shuttle heat transfer as a function of coldend temperature with the hot end at 300 K.

Figure 2 shows the speed dependence of the shuttle heat transfer. This figure shows how

•1/2

the shuttle heat transfer gradually varies from a n

' dependence at low speeds to an independence of n at higher speeds. The experimental results agree fairly well with calculated

values for both He and N« gas in the gap.

in •0

02-

0.01

"I I I I 1 I IT] 1

1—I Mill

He Gas Limit

1

Stroke=2 45 cm :

"J £ lExperm^Ul —Calculated

_i—i i i i i il 5 10 20 SPEED, rpm

50 100

Figure 2. Measured and calculated shuttle heat transfer as a function of displacer speed for a cold-end temperature of 140 K.

50
We conclude that at least for the parameters used in this experiment, shuttle heat

transfer can be estimated to within about 15% by combining the two expressions given by

Eqs. (2) and (5) in the manner prescribed by Eq. (6). This imples the two thermal resistances are in series and that the necessary simplifying assumptions can be met in practice.

Optimization of the parameters of the Stirling refrigerator can then be carried out analytically. In the tests done here the shuttle heat transfer was about five times the static

conduction for a stroke of 1.19 cm. These results show that a more optimum design would have

a shorter stroke and larger diameter to minimize the total heat transfer. Because k and C

3 are proportional to T at lower temperatures, the calculated values may not be in quite so good agreement with experiment at such temperatures unless Eq. (5) 1s Integrated properly

between T. and T . Measurements down to about 5-10 K would be useful and are in progress.

REFERENCES

[1] J.E. Zimmerman and Ray Radebaugh, this publication.

[2] F.J. Zimmerman and R.C. Longsworth, Adv. Cryog. Eng. 1_6, 342 (1970).

[3] J.B. Harness and P.E.L. Neumann, Proc. Int'l Cryog. Eng. Conf., 5th (1972) p 97.

[4] P.J. Schneider, Conduction Heat Transfer , Addison-Wesley, Cambridge, MA (1955) p 276.

[5] T. Ashworth, J.E. Loomer and M.M. Kreitman, Adv. Cryog. Eng. 1_8, 271 (1973).

[6] W. Reese and J.E. Tucker, J. Chem. Phys. 43, 105 (1965).

[7] B. Wunderlich, Rensselaer Polytechnic Institute Report No. TR-17, Office of Naval Re- search Contract Nonr-591(19) (1968).

[8] J. A. Barclay, S.S. Rosenblum and W.A. Steyert, Cryogenics 16, 539 (1976).

51
XV International Congress of Refrigeration, Venice, Italy (September, 1979).

ANALYSIS OF REGENERATOR INEFFICIENCY FOR STIRLING-CYCLE * REFRIGERATORS WITH PLASTIC DISPLACERS

Ray Radebaugh

Thermophysical Properties Division

National Bureau of Standards

Boulder, Colorado, U.S.A.

1. INTRODUCTION

One of the major sources of heat flow to the cold end of a Stirling-cycle re- frigerator operating below about 15 K is due to regenerator inefficiency. Many

analyses and experiments have been performed previously, but all have dealt with

the case of negligible thermal resistance internal to the regenerator material.

Recently temperatures as low as 3 K have been reached with a Stirling-cycle refrigerator that uses fiberglass-epoxy for all the low temeprature parts. Ill Plastic

regenerators - displacers are necessary for operation of SQUID magnetometers, but

the internal thermal resistance of the plastic dominates the behavior of the regenerator. This paper analyzes the inefficiency of such regenerators and compares

the results with experiment. In addition the effect of heat flow to the cold end caused by non-ideal gas effects ere considered.

2. DEVELOPMENT OF MODEL

In all of the refrigerators built to date with plastic displacer - regenerators, the regenerator is the annular gap between the displacer and the walls.

For the 3 K refrigerator an additional annular gap outside the first cylinder was

provided /I/. The model we use to describe the regenerator behavior is one of a seal- infinite solid with the surface undergoing a cyclic temperature variation.

In addition there is a temperature gradient parallel to the surface. Nusselt 111 has considered the case of regenerators which have a finite conductivity normal to the gas flow and zero conductivity parallel to the gas flow. Of the cases considered by Nusselt this one is the closest to our situation except that the thermal conductivity and specific heat are strong functions of temperature. The complete

set of second-order partial differential equations then become so complex that their solution would require considerable computer time.

In order to simplify the equations governing the regenerator behavior we make

the following assumptions:

(1) temperature variation at the surface is sinusoidal,

(2) surface temperature is the gas temperature,

(3) gas volume in regenerator is negligible, compared with expansion

volume,

(4) the sinusoidal heat flow is only in the radial direction,

(5) the expansion and compression space are isothermal,

i.e., the Schmidt analysis 131 can be used,

(6) various heat flows to the expansion space are independent

of each other. Assumptions 2-4 appear to be quite reasonable in the range of the present experiments. Assumption 1 is made since it greatly simplifies the calculations even though the mass flow of gas through the regenerator is only approximately sinusoidal. For a sinusoidal temperature variation at the surface, the heat flow into

(or out of) the regenerator during one half the cycle is /4/

Q - 2AAT /kC„/2Trn . (1) o v

For an annular gap heat can flow both outward and inward in the solid material and

so the surface area A should be 2irDL. The amplitude of the temperature variation

52
decreases with the distance from the surface. The thermal penetration depth X is defined as

X - /k/C *n v(2)

where the amplitude is 1/e of that at the surface. Thus the model considered is valid only for wall thickness, d > X. The validity of assumption 2 and of using the semi-infinite solid model for

d - 3 mm is shown in Fig. 1 which compares the AT of the surface of the seml10 >

CM \~3

:h

10"=-

io2 k

10

= 1 1 1 1

| 1 1 I 1 1 1 1 1

| 1 11 =

T = 5K ~

- G-10 THERMAL PENETRATION^.

"

Z" ^^^--*"~" G-10 3mm WALL THICKNESSX = —5*""" ^ -

- LEAD 3mm WALL\ _ ^_ \ _ —<^

' ^^ TURBULENT FLOW^

2 ^^LAMINAR FLOW "= - .^""HELIUM GAP = 0.1mm 2

=- ^_^—-—^^LEAO THERMAL PENETRATION -

-3

..!.. I i i 1 i i I i i I i i 1 i tl

0.5 1.0 10 50 SPEED, Hz

Fig. 1. A comparison of the AT times area per unit heat flow as a function

of speed in the gas and in the solid regenerator.

Infinite solid with that due to a wall 3 mm thick which has X « » and with the AT across the helium gas. For speeds of 1 & and above the dominant AT is due to the finite thermal penetration depth of the G-10 fiberglass epoxy. The thermal penetration depth Increases rapidly as the temperature is lowered and at 5 K is X = 2.1 mm

for n - 1 Hz. Heat flow into the regenerator comes from decreasing the temperature of the helium gas. The amplitude AT is found by equating the heat flow from the gas, SdT, with that in the solid, i.e., Eq. (1). After some rearranging of terms, we

AT mC (dT/dx) /4D/irkC ,/2n P v(3)

Because of assumption 4 we take dT/dx to be that caused by conduction heat transfer. An iterative procedure is required to find AT since the average temperature used

to calculate dT/dx depends on AT . At the cold end the regenerator temperature os- cillates between T and T + 2AT .° The calculated ATQ usually is less than about 0.7 K for T » 5 K. The incoming gas then enters the expansion space at an average

temperature of 1 + AT and thus causes an effective heat flow of

" naC AT xeg p c n(mC ) (dT/dx)/4D/rrkC /2n

p v

(4)

53
The term mC Is found from an integration over the crank angle using the Schmidt

analysis. p The overall refrigerator behavior is calculated by using the heat balance equation

<ref \et +

^rad + %

+ Q

. + Qfrict +

<>reg + % (5)

The shuttle heat flow, Q has been calculated previously 15/ for plastic displacers. The term Q occurs in a Son-ideal gas because the total enthalpy flow from T to T

is greater than enthalpy flow in the reverse # direct ion due to the pressure variations.

To calculate the minimum temperature we set Qnet " °* '^ne weakest assumption is 5, since X for helium at O.lMPa is only 0.2 mm at 5 K, but it increases to 7.7 mm at 300 K. Thus nearly adiabatic behavior occurs at 5 K and the Schmidt analysis for Qref would be too high by possibly a factor,or two. ^ Since the Schmidt analysis is also used to find the mass flow, the terms Qre_ and Qh would also be t<jo high by

roughly the same factor. For Q t and T <°7 K the terms Qr and Qy dominate

the others so assumption 5 may be satisfactory when calculating the minimum temperature.

3. COMPARISON WITH EXPERIMENT

Calculated behavior is compared with the experimental behavior of the re- frigerator described by Zimmerman and Sullivan 111 . The calculations for the minimum temperature were done by setting Qnet " and calculating the required

regenerator length for different expansion apace temperatures. Because of the isothermal assumption the calculations for Qjj cannot be done in the two fluid

region reached in the experiments. The term % increases rapidly as liquid is formed but only 32 liquid was produced in the experiments. Thus extrapolation to the lower temperatures should be reasonable. Figure 2 shows calculated regenerator

20

1 1 T 1

T

u=12

1 1

.OK

-

10

5

2

o o*\

£ \ \ 4 ~—'—^.\ ,9kPa I, ^ ">-C^ Y

-\ v^ * \ - x \^s^-19kPa } \ 47kPa^\^S>,,

Pmean!86kPa

- \

1 _ Vs.skX^ 80kPa — '\ -

.5 i 1 1 1 \ 1

-

4 5 6 7 8 9 10 TEMPERATURE, K

Fig. 2. Calculated regenerator lengths as a function of the expansion space

temperature. All calculated curves are for Tu

- 12 K except the lowest curve. The experimental points at L - 12 cm are for Tu

- 8.6

12 K. P„._ - 80kPa.

54
8.6 lengths for different expansion space temperatures and pressures. With Tu

K, P -27 kPa, and L - 12 cm the experimental result of 3.1 K is consistent

withm

the extrapolated value from calculations. For Tu - 12 K, Pmean 80 kPa, and

L - 12 cm the calculated value of 4,4 K is in good agreement with the experimental

value of 4.1 K. The optimum pressure for T - 4 K and Tu

- 12 K is about 27 kPa

according to Fig. 2. The optimum diameter and stroke are about 6 mm and 10 mm

respectively for a displaced cold volume of 0.3 cm . Figure 3 shows a comparison of Qnec

for experiment and theory. The minimus

temperatures agree well but the agreement in Qnet deteriorates as Qnet increases,

presumably due to the adiabatic behavior. Figure 4 shows the magnitude of the various heat flow terms. Note that Qreg a n P from Eq. (4), since m a P. Figure 4 shows only a slight deviation from P

2 behavior becaus^ of the pressure dependence of Cp

. For Cp independent of temperature Qreg a T for T = 5 K.

30

25-

20-

10

5 -

Pmean=

80 kPa

T

u

= 8.7 K

' O EXPERIMENT

± j_

4 5 6 7 TEMPERATURE, K

Fig. 3. A comparison of the theoretical and experimental net refrigeration

powers as a function of temperature.

4. DISCUSSION

• . • 3/2_2 Because Qre£ a nP and Qreg a n T> we see immediately the advantage of using

low pressures and low speeds to improve the overall performance of the low temperature stage where Qreg becomes large. Further improvements in the regenerator should be directed toward Increasing Cy but keeping k small

to prevent an increase in Qcond - The term Q^ is independent of the regenerator

length and so it can easily dominate ^ the heat flow for long regenerators. Fig. 4 shows that Qh is always larger than Qreg for L - 12 cm and that Qg has a similar pressure ( dependence. The refrigerator performance should improve if He were used because Qg would be reduced. The model discussed here appears to predict the minimum temperature to within about 10Z. Thus for Qnet = ° tnis model c*° De used

to optimize the design of such refrigerators.

55
20

5 -

0.5

0.2

0.1

T I I I I I I I 1 1 I I I I I

J »'''

Fig. 4. Behavior of the various heat flow terms as a function of pressure

for L - 12 cm.

<frict

;net ?rad

?ref reg

NOMENCLATURE

surface area for heat transfer

specific heat of helium at constant pressure volumetric specific heat of the regenerator material

diameter of the annular gap in the regenerator

thermal conductivity of the regenerator material

length of the regenerator

total helium mass flow into the expansion space operating speed of refrigerator pressure of helium gas cumulative heat transfer

heat transfer rate to the expansion space due to conduction

heat transfer rate to the expansion space due to friction

heat transfer rate to the expansion space due to enthalpy

imbalance* net heat transfer rate to the expansion space heat transfer rate to the expansion space due to radiation

total refrigeration rate in the expansion space effective heat transfer rate to the expansion space due to regener- ator inefficiency

56
Q_ heat transfer rate to the expansion space due to shuttle transfer

/5/ T temperature of the expansion space T

u temperature of the hot end of the regenerator AT

Q half-amplitude temperature variation of the regenerator surface dT/dx temperature gradient in the regenerator in the direction of gas flow

A thermal penetration depth due to surface temperature variations

REFERENCES

*Work supported by U.S. Naval Air Systems Command. Contribution of the U.S. Government, not subject to copyright.

1. J. E. Zimmerman and D. B. Sullivan, "A Milliwatt Stirling Cryocooler for Temperatures Below 4 K", Cryogenics, March (1979), to be published; D. B. Sullivan

and J. E. Zimmerman, "Very-Low-Power Stirling Cryocoolers Using Plastic and Composite Materials", these proceedings, and references cited therein.

2. W. Nusselt, Z. Ver deut. Ing. 71, 85 (1927), and M. Jakob, Heat Transfer, vol.

II. J. Wiley & Sons, New York (1957), p. 288.

3. C. Walker, Stirling-Cycle Machines, Clarendon Press, Oxford (1973), p. 24.

4. P. J. Schneider, Conduction Heat Transfer, Addison-Wesley, Cambridge, MA

(1955), p. 280.

5. R. Radebaugh and J. E. Zimmerman, "Shuttle Heat Transfer in Plastic Displacers

at Low Speeds", NBS Special Publication 508 (1978), p. 67. Available from U.S. Government Printing Office, Washington, D.C. 20434.

ANALYSE DE L'INEFFICACITE DU REFENERATEUR DES REFROIDISSEURS A CYCLE-STIRLING AVEC DEPLACEUR EN PLASTIQUE

SOMHAIRE: La source majeure de trans ferts therm iques dans un refroidisseur a cycle-Stirling fonctionnant a moins de 15K environ, provient de l'inefficacite" du rdgenerateur. Toutes lea experiences et analyses precedentes de l'inefficacite' ont 6t6 effect uees dana le caa d'une resistance thermique interne negligeable du materiau

du rfgenerateur. Dans le cas du refroidisaement d'un magnetometre SQUID, les organea baaae-temperature du refroidisseur sont en plastique. Done, la resistance

thermique interne eat la resistance dominants. Cet expose examine le transfer t ther- mique <JQ a l'inefficacite du regenerateur loraque les materiaux sont a basse diffusion thermique. Dans un tel cas, le transfert de chaleur eat proportionnel a

n 3/2 p2 ( od n e8 t ia vitesse et P la pression de l'helium gazeux. Du fait que

le taux de refroidissement est proportionnel a nP, une reduction de la vitesse et ce la pres8ion r6duit la temperature minimum jusqu' a ce que le transfert thermique

de va-et-vient, proportionnnel a rr'^, et le transfert thermique de conduction,

deviennent importants. Une comparaison avec lea result at s experiment eux a des temperatures de l'ordre de 3K, sera fnurnie.

57
Cryogenics 19 (1979), pps. 170-171

A milliwatt Stirling cryocooler for temperatures below 4 K

J.E. Zimmerman and D.B. Sullivan

We have built a single-stage, twin-piston Stilling machine'

producing several mW of refrigeration at temperatures in the range of 3 to 4 K, and rejecting 50 to 1 00 mW at temperatures of 8 to 14 K. The working fluid is helium at subatmospheric average pressure. During operation, several mm3 of liquid helium collects in the dead space at the cold

end, and is present throughout the operating cycle. Thus

the operating temperature is essentially the equilibrium vapour

pressure of helium at the average pressure - for example,

3.1 K was achieved with an average pressure of 0.027 MPa.

and 4.0 K with an average pressure of 0.08 MPa. The hot end temperatures were 8.6 K and 12 K respectively. The cold parts of the unit shown in Fig. 1 consist of two

concentric pistons and a stationary outer cylinder, all three members being made of commercially available spun-glass epoxy 'G-10' rod and tubing requiring no machining by us. Small radial clearances and large length-to-diameter ratios prevent significant gas leakage between the ends of the pistons

at room temperature. Thus, the working fluid provides gas

lubrication between adjacent surfaces, and no solid sliding

seals (ie piston rings) are required. The two pistons are driven 90° out of phase, with a stroke of 1 cm. by

connecting rods and a crankshaft (not shown) at room

temperature. The latter are enclosed in a metal housing (~ I t volume) which is filled with helium gas to the desired pressure. The only moving seal required is an

O-ring around the crankshaft end which extends outside

the housing to an electric drive motor. Owing to the reduced

viscosity of helium at low temperature, there should be

little pressure drop across the regenerator gaps so the

pressures in the expansion and compression spaces should

be essentially equal at all times. In principle the motor could

also be enclosed in the metal housing and no seals at all would

be necessary. Since there are no sliding seals on the pistons, the average pressure in the cylinders is identical to that in the upper housing. Most of our data were taken with an

operating speed of 1 Hz. For the experiments reported in this note we maintained the hot flange on the upper cylinder at temperatures in the range of 8 to 14 K by means of a weak thermal link to a liquid helium bath and an electrical heater mounted on the hot

flange. The temperature of the hot flange was controlled simply by varying the heater power. The boil-off rate from

the helium bath was used as a measure of the heat input to the system and, in particular, we used changes in the boil-off

rate to estimate the heat ejected from the hot end of the

refrigerator during operation.

The authors are at the Electromagnetic Technology Division, National Bureau of Standards, Boulder, CO 80303. USA. Received 9 January 1979.

As shown in Fig. 1, we incorporated a pair of universal

joints in the central piston rod at the level of the compression space. This allows the lower part of the piston to centre itself in the cylinder. Without this feature, slight misalignment of the cylinder surfaces above and below the compression space caused those surfaces to bear against the

sides of the piston. The associated frictional heating degraded the expected performance considerably. Another feature, which may be important, shown in Fie. I

is the 'stationary insert' around the lower piston. This liiseri

is a piece of the same material as the upper piston. Having

equal annular gaps at both the inner and outer surfaces of

the insert provides two parallel paths for fluid (low between

the expansion and compression spaces, giving a total area

To room temperoture

connecting rods and cronkshaft

W>/>////S//A

Outer cylinder

16 mm od

Clearances -30>»m

. Upper piston (onnulor) -12 mm od

Universol joints

-Hot end flange

- Compression space

-Lower piston -6mm didmeter

Stationary insert

- Regenerator gops ~30^m

Eipantian space

.Cold end

Fig. 1 Sketch of single-stage Stirling cryo-cooler with approximate

dimensions

58
Fig. 2 Temperature vs time of the cold end with the cryocooler stopped at maximum expansion (lower piston 'up ) bnd minimum

expansion (lower piston 'down')

for regenerative heal exchange three times the area of the inner gap alone. At the 1 Hz operating speed, we estimate

that heat diffuses into the material less than 1 mm, so

effectively the insert provides three times the regenerator

heat capacity. In order to analyse the effectiveness of this particular feature, one needs to know the heat capacity and

thermal conductivity of the material, which we have measured

and will report separately. With the upper end temperature at 8.7 K, the refrigerator maintained a temperature of 4 K at an operating pressure of

0.08 MPa. but there was some reserve cooling capacity. The

operating temperature was limited by the pressure of the

liquid phase. Heat was added at the bottom end to determine the reserve cooling capacity. The system maintained

liquid helium at 4.0 K up to a heat input of 4 mW at which

point the system reverted to single phase gas operation. At

9 mW input power the temperature rose to 5.5 K. The

Carnot coefficient for the cooler is 4/(8.7-4) = 0.85. From

changes in the helium boil-off rate, we estimated that the

heat ejected at the hot end is about 50 mW. Thus the actual

refrigeration of 4 mW at 4 K is a reasonable fraction of the

theoretical cooling power (50 mW x 0.85 = 42.5 mW). A calculation of the work input yields a number of the same

order as the heat ejected, but not much information can be

elicited from this since the two phase nature of the

refrigerator operation complicates the analysis.

Fig. 2 shows two separate warming curves for the cold end

of the refrigerator, recorded immediately after the drive motor was switched off. These demonstrate the presence

of liquid helium in the system. The temperature rises gradually

at first since the evaporation of the liquid increases the

overall system pressure. After several minutes the temperature

increases more rapidly, indicating evaporation of the last

bit of liquid. The total quantity of liquid helium present

in the system can be estimated from the pressure change

during the slow portion of warm-up of the bottom end and comes out to be — 0.02 cm

3

. The two curves correspond to stopping the refrigeration cycle at two different positions separated by 1 80°. The first portions of the curves are similar

in shape although the temperatures are different. The time

interval required for all the liquid to evaporate varies because a different quantity of liquid is present during each

part of the cycle. The variation in slope beyond the break

point indicates different heat capacities at the bottom end.

that is, if the bottom piston is in its uppermost position then

the heat capacity of the extra vapour reduces the warming

rate. This is a clear indication that the heat capacity of the vapour in the expansion space is significantly greater than

that of the G-10 epoxy. the implication being that the additional channel for regenerative heat exchange provided by the stationary inset may be an essential feature. We believe this is the first time that a Stirling machine

operating as a refrigerator has been shown to liquefy the working tluid (helium or any other gas), and the temperature

achieved is about a factor of two lower than has been reporled

previously for small Stirling. Gi fiord- McMahon. and

Vuilleumier machines. 3 A machine of this t\ pe appears to be a practical alternative to a Joule-Thomson expansion system for the final stage of a small helium liquefier or cryocooler. The heat load that our particular machine would impose on

the higher temperature stage (at 8 to 14 K. for example), is low enough that the total system should require an order of magnitude lower mechanical power input than any

existing small cryocooler.

This work was supported by the Naval Air Systems Command

References

1 Walker, Ci Slirlinp-Cvclc Machines Oxford University Pross (1973)

2 Zimmerman. J.E., Rynn. T.M. NBS Special Publication SP-508. National Bureau of Standards (1977)

59
"Future Trends in Superconductive Electronics," (B. S. Deaver, Jr., C. M. Falco, J. Harris, and S.

W. Wolf, eds.) AIP Conference Proceedings Number 44, New York, American Institute of Physics (1978),

pps. 412-420.

CRYOCOOLERS FOR SUPERCONDUCTIVE ELECTRONICS*

J. E. Zimmerman

Cryogenics Division

National Bureau of Standards

Boulder, Colorado 80303

ABSTRACT

Superconductive electronic devices are generally operated and

will continue to be operated in liquid-helium cryostats. Appropriate

small closed-cycle cryocoolers for these devices do not exist.

Development of cryocoolers, including efficient use of conventional

gas refrigeration cycles and innovative thoughts on new and unconventional methods, could greatly enhance the usefulness of small-scale

superconductive electronics. Large cryocoolers are already welldeveloped and can be adapted to the requirements of large-scale

systems, such as computers, requiring a watt or more of refrigeration.

INTRODUCTION

In this paper I will refer repeatedly to two sources

1

* 2

. Ref.

1 is the proceedings of a 1977 conference on "Applications of Closedcycle Cryocoolers to Small Superconducting Devices," and is publicly

available. Ref. 2 is the proceedings of a 1973 conference on "ClosedCycle Cryocoolers," and is available for "limited distribution." A

third significant source of information and speculation is the

proceedings of the 1975 Navy Study on Superconductive Electronics. 3 Other useful and readily available references are the proceedings of

the 1976 SQUID conference" and the proceedings of the 1976 NATO

study on small superconducting devices 5.

This author's demonstrated qualifications as an authority on cryocoolers consist of his having built (1) in 1949, a Simon singleexpansion helium liquefier, a mechanism which at the same time was made obsolete by the Collins Helium Cryostat, (2) a few years later,

a cyclic magnetic refrigerator for temperatures down to 0.1K, a machine which was immediately replaced by the helium dilution refrigerator, and (3) recently, a laboratory-model Stirling cryocooler

specially adapted to operate a SQUID magnetometer. The trend established by the earlier work, probably has been reversed by the recent

achievement, since Stirling machines are now being extensively developed both for pumping heat and for doing work, and this conference

is evidence that the SQUID and other superconductive devices are seen

as having great future potential. In fact, it should also be pointed

Contribution of the National Bureau of Standards

Not subject to copyright.

Support in part by the Office of Naval Research

and the Naval Air Systems Command

60
out that cyclic magnetic cooling systems are currently being developed by Steyert and others for temperatures between 1 and 300K, and it

is suggested below that the single-expansion process of liquefying

helium may still be of some use in certain systems (p. 81 and p. 49,

ref. 1).

Very few closed-cycle cryocoolers are now in use, or have ever been used, to operate superconducting devices. The reason for this

became evident in 1947 or shortly thereafter, when the Collins Helium

Cryostat became generally available. This machine was advertised and

sold as a "cryostat", a device for maintaining a constant low temperature. Very quickly it became obvious that the machine was more

cost-effective if used not as a cryostat but as a liquefier, for

servicing any number of separate small liquid-helium evaporative

cryostats.

ESSENTIAL HEAT INPUTS INTO CRYOSTATS

With only a few isolated exceptions, (e.g. Hartwig. p. 135, ref.

1) liquid-helium cryostats have been used for all research in superconductivity and superconductive devices up to the present. Now,

however, the beginning of change is apparent. As superconducting

devices are found to be useful for a variety of purposes unrelated

to low-temperature physics, users are searching for ways of reducing

the cost and inconvenience of liquid-helium cryogenics. One approach

is to make dewars larger and more efficient, with hold times of the

order of 10 3 to 10

1* hours (e.g. the 180-liter VLF SQUID cryostat of

Dinger et al. , with a hold time of 100 days 6

, the 800-liter Infrared

Astronomy Satellite (IRAS), with a projected hold time of one year , and a similar 800-liter satellite for the Stanford relativity experiment, projected to operate one year in orbit at 2K8

. More indicative

of the nature of things to come, we hope, are the systems in which a cryocooler is incorporated into a medium-sized liquid-helium dewar to cool one or more radiation shields and greatly reduce the evaporation

rate. Temperatures well below 20K can be maintained with available

small commercial coolers. Black-body radiation at 20K is only 9.1

mW/m2

, and if this figure can be reduced by a factor of 100 by the use of highly reflecting surfaces, giving 91uW/m2

, the corresponding

evaporation rate due to radiated heat is 0.0031 Jl/day-m2. A spherical

container with a surface area of lm2 would hold 95 liters, giving a time of 3 x lO

1* days (80 years) for complete evaporation. In most

practical cases, with a radiation shield at 20K and a helium bath at

4K, the main heat input will be conduction along mechanical supports

and electrical leads, but even so, the heat leak can in principle be

made very small. If a 95-liter container is supported by spun-glass

cords, 10 cm long, calculated to withstand 100 G loading in

three orthogonal directions (six cords) , and assuming a total

61
mass twice that of the helium itself, the heat leak is calculated to be 2mW, giving a hold time somewhat less than 4 years (40 years for

10G loading). The enthalpy change of helium between 4 and 20K is

about 4 times the heat of vaporization, so if the vapor can be used to cool the supports, this evaporation rate might be reduced considerably, at the expense of geometric complexity. Size of electrical

leads depends upon the particular devices in the cryostat. Let us consider (as we often do) an rf SQUID and suppose that the rf line to it is a thin-film stripline with a parallel dc resistance of 0.1ft between 4 and 20K. Invoking the Wledemann-Franz law leads to a heat

leak of about 45 WW, which is smaller than any of the other heat

leaks calculated above.

This exercise in arithmetic is highly idealized, but it does

indicate the possibility of greatly reducing the evaporation rate of

liquid helium by using a cryocooler in a carefully-designed system.

A 70-liter cryostat incorporating a cryocooler to maintain the inner

radiation shield at 15K has been built by a commercial firm (W.

Goree, private communiation) . This system demonstrated a hold time

greater than one year (0.19£./day evaporation rate). The heat leak

into the helium, about 5.5 mW, was calculated to be due almost en- tirely to the 75 mm-diameter stainless steel neck tube, which provided easy access to the helium bath and also served as mechanical

support for the reservoir.

Conventional liquid-helium cryostats, such as those mentioned

above, without liquid-nitrogen jackets, typically evaporate from one

to several liters per day. One liter per day indicates a heat input

into the liquid of 30 mW. The total enthalpy change of the vapor

between 4 and 300K is roughly 70 times the heat of vaporization. The

cooling capacity of the vapor, proportional to the specific heat Cp,

is nearly independent of temperature over most of the range. This is

not, in general, the optimum distribution of cooling capacity vs. temperature. A closed-cycle cryocooler with the optimum distribution

is therefore, in principle, more efficient than an evaporative cryostat. Related to this is the important point that, for many small

low-power superconducting devices, the lion's share of the cooling

capacity is required at higher temperatures to intercept heat radiated and conducted from room temperature. To specify the performance

of a cryocooler only in terms of its cooling capacity at the cold

end, as is commonly done (e.g. p.iv, ref . 3), is therefore misleading.

CLOSED-CYCLE CRYOCOOLERS VS EVAPORATIVE CRYOSTATS

The relative merits of closed-cycle cryocoolers and evaporative

cryostats have been debated many times. In 1966, Winters and Snow

showed that an evaporative cryostat is economically preferable to a

62
cryocooler for a heat load of 100 W (3300 8,/day evaporation rate) if

the duty cycle is less than 95 hours per year for 10 years

9

. If one may extrapolate from this by 4 or 5 orders of magnitude, it might be

concluded that a cryocooler could not possibly be economical at milliwatt refrigeration levels. Nevertheless, economic analyses are imperfect at best, and may change with time as the data base changes.

One such change now occurring is that superconductive electronics is

becoming a widely used technology with large duty cycles. Another

change which one may expect in the future is that helium gas will

become more expensive as the supply runs out. There is also a possibility that inventive genius may change the economics of small

cryocoolers by discovering better and simpler refrigeration methods

(see for example W. A. Little, this conference, also R. Radebaugh,

p. 7, W. A. Little, p. 74, W. A. Steyert, p. 81, R. Radebaugh, p. 93,

and J. E. Zimmerman and R. Radebaugh, p. 49, ref. 1). A factor which

might significantly affect the economics of cryocoolers is the

development of high-Tc devices (M. Beasley, this conference and p. 167, ref. 1).

COST AND RELIABILITY OF CRYOCOOLERS

Data on cost and physical characteristics of practically all

cryocoolers built up to 1973 have been collected by Strobridge 10

. More data on small cryocoolers, primarily for infrared sensors and

space applications at temperatures of 10 to 100K, are given in ref- erence 2 (W.S. Sims, -p. 1, V. C. Johnson, p. 11, and W. D. Clarke

III, p. 17). This reference provides interesting information (p. 15)

such as that the Air Force had expended over 9 x 10 6 dollars, as of

1973, on two objectives (1) "decrease the cost of ownership for

airborne cryogenic coolers, and (2) "increase the unattended life of

cryogenic coolers for spacecraft to something in excess of three

years." The second objective had not, and probably has not yet, been

achieved. The reference contains only one paper relating to cryocooler technology for elemental superconductors (W. Lawless, p 417,

ref. 2, see also R. Radebaugh, p. 7 and p. 93, ref. 1) but several

papers (e.g. B. Renyer, p 147, ref. 2) describe cryocoolers capable

of cooling high-T devices. Reference 2 also contains some speculation on cost of cryocoolers. S. Horn e_t a_l. , p. 73, and W. Sims, p.

1, ref. 2, have estimated prices of 1000 to 2000 dollars for the very

simplest (Vuilleumier) single-stage miniature 77 K closed-cycle

cryocoolers, in lots of 1000. Miniature multistage Stirling and

Vuilleumier machines for low temperatures, 6 to 8K, should not nec- essarily cost much more than this, since, by using the steppeddisplacer principle of Fokker and Kohler 11

, the number of essential

moving parts (approximately 5) is precisely the same as for a singlestage machine, and only the displacer geometry is more complicated.

Other information of general interest from reference 2 are the

comments on reliability of closed-cycle cryocoolers. W.Clarke III,

p. 17, stated that "failures occur between 120 and 300 hours of

63
cooler operating time on all Air Force airborne systems," which includes Joule-Thomson, Stirling, and Gifford-McMahon cryocoolers.

Chellis, p. Ill, ref. 2, reports a mean time between failure of 20000

hours for several hundred Gifford-McMahon cryocoolers used (on the

ground) to cool parametric amplifiers for satellite communications

and also for military purposes. Clarke and Chellis were reporting on different systems with different reliability requirements, but the

contrast is nevertheless impressive. More recent information on reliability is given by Chellis (p. 109, ref. 1) on 20K machines and

by Higa and Wiebe (p. 99, ref. 1) on 4K machines. About 30 or 40 of

the latter are in use to cool masers for deep-space probes.

LARGE AND SMALL MACHINES AND APPLICATIONS

In the literature and in private conversation, cryocoolers with

one to several watts of refrigeration at the cold end are often

called "small" or "miniature." In the context of this paper, such

machines will be referred to as "large." This paper will deal primarily with concepts of cryocoolers for devices such as single

Josephson junctions or small arrays of junctions whose inherent

refrigeration requirements are in the sub-milliwatt level. At least one author (W. A. Little, p. 74, ref. 1) has pointed out that present

commercial cryocoolers are grossly mismatched to such devices. With

regard to large superconducting systems such as computers, which may

actually require a watt or or more of refrigeration, present cryocooler technology (F. Chellis, p. 109, R. Longsworth, p. 45, W. Higa,

p. 99, ref. 1) should be readily adaptable, perhaps with some effort

to reduce vibration levels (see for example, J. Cox and S. Wolf,

p. 123, ref.l). The computer application of cryocoolers was considered in detail by van de Hoeven and Anacker (p. 221, ref.l), who indicate that refrigeration requirements may fall in the range of 0.1 to 10W.

A few "large" machines have been used for superconducting devices

and research. The first such use was by W. Hartwig and his colleagues to cool a superconducting Pb cavity (for a review of this and

other uses see p. 135, ref. 1). Operation of a SQUID susceptometer is

described by D. Vincent (p. 131, ref. 1), and a cryocooler has been

used or several years for low-temperature research by W. Little

(private communication) . Both of the latter two machines were originally provided by the Office of Naval Research (E. Edelsack) for

operational evaluation with superconducting devices.

It is arguable that there is a need for small cryocoolers for

such applications as SQUID magnetometers for geothermal prospecting

(W.D. Stanley, p. 205, ref. 1, and J. Clarke, ref. 5), gradiometers

for magnetic anomaly detection 12 and for biomagnetic research

(reviewed by S. Williamson and L. Kaufmann, p. 177, ref. 1),

64
Josephson voltage standards (L. Holdeman and C. Chang, p. 243, ref.

1), microwave and millimeter-wave detectors (J. Edrich, p. 159 and M. Nisenoff, p. 221, ref. 1) and a variety of other laboratory research

instruments. There have been virtually no studies of small or miniature cryocoolers other than the recent one by W. A. Little on microminiaturizing the Joule-Thomson process (p. 74, ref. 1).

As far as I am aware, the only experimental small (milliwatt)

cryocooler work currently going on is that by me and my colleagues on

a low-speed low-power 4-stage split Stirling machine for operating a Nb SQUID at 8.5K (p. 59, ref. 1). This machine is unique in several

respects, (1) it requires an order- of- magnitude- lower input power

than comparable machines that have been built for the same temperature

range, (2) it employs a gap regenerator, making a displacer and heat

exchange mechanism of ultimate simplicity, and most important (3)

nylon and epoxy-glass composite materials are used for the displacer

and cylinder to minimize magnetic interference caused by moving parts near the low-temperature end.

The requirements for small cryocoolers were estimated by

M. Simmonds (p. 207, ref. 1). Except for the susceptometer, his

estimates of refrigeration requirements for various devices are all

of the order of a few mW. Even these estimates are considerably

higher than the above calculations would indicate, but it may be

assumed that he is biased more toward commercial realism than aca- demic idealism. In any case, even if extremely low refrigeration

power satisfies the steady-state requirements, cooldown times may be

too long for some applications. Estimates of acceptable cooldown

times and other parameters vary considerably (M. Simmonds, p. 207 and

M. Nisenoff, p. 221, ref. 1).

FUTURE TRENDS

With regard to the future of superconductive electronics, there

is a great need for intelligent but uninhibited analysis and experimentation on cooling methods for small superconducting devices and

systems. Classical gas-cooling systems like Stirling and GiffordMcMahcn machines are highly developed and effective for temperatures

as low as 6 to 8K. This is tantalizingly close to the temperature

range needed for most present devices. It is quite adequate for

high-T devices, and although the latter will undoubtedly serve some

of the needs of superconductive electronics, lower-temperature cryocoolers will be essential for lower noise devices, for devices like

the super-Schottky diode 13 which must operate well below the transition temperature, and to make use of elemental superconductors with

various special properties. To span the temperature region from

65
about 8K down to 1 or 2K, the development of solid-state processes as alternatives to Joule-Thomson or other processes requiring compressors

would be highly desirable. On the other hand, perhaps a low-temperature low-pressure stage using helium (or helium 3) gas in a Stirling

cycle would be practical. For very low refrigeration power, a few

cm

3 of liquid helium produced by a single expansion, with a starting

and shield temperature of 8K, could last practically forever. Longterm reliability is probably the most difficult thing to achieve with

mechanical systems, so we should emulate Steyert by considering other

refrigeration processes for higher temperatures as well. Even

though conventional gas refrigerators may be hard to beat, there are numerous ways to improve and simplify these systems—gas bearings,

thermal compressors, absorption-desorption compressors, resonant

spring-mounted or free pistons and displacers driven by mechanical

reactive coupling rather than by crankshafts and connecting rods, and

(no doubt) hundreds of other good ideas.

Radebaugh (p. 7, ref. 1) has written an unusually informative and

stimulating review of refrigeration fundamentals. In addition to the

commonly-used gas and gas-liquid systems, he describes a broad range

of other systems, tells what kind of forces (electric, magnetic,

tensile, centrifugal, etc.) are required to change the entropy, and

estimates the minimum useful entropy change. An example of the many

less-well-known systems is a rubber-band refrigerator, with which a modest amount of refrigeration might be obtained in the neighborhood

of room temperature.

Space applications of superconductive electronics pose special

problems for cryocoolers in terms of weight, power and reliability.

Work on large cryocoolers (mostly classified) under the sponsorship

of Wright-Patterson Air Force Base is continuing, with a multi-stage

Vuilleumier cryocooler under development for launching in a spacecraft within a few years. For certain farout missions, cooling of

superconducting devices by radiation into space has been suggested.

I have estimated that a black aluminum sheet 1 mm thick, carefully

shielded from the sun and not too close to any of the planets or their satellites, and attached to the spacecraft by a practical lowconductivity suspension, should reach a temperature of 5 to 10K in

about 2 days. The estimate is based upon evidence that the total

energy flux in interplanetary space, including zodiacal light, starlight, and cosmic background, but not including sunlight, adds up to something like 10 to 100 uW/m2

. Solar energy flux at the earth's

orbit is 1.3 kW/m2

. A number of applications of superconductive and

other devices in space were reviewed by E. Tward and P. Mason and by

J. Vorreiter and C. McCreight (p. 227, and p. 153, ref. 1).

66
As for specific suggestions, there is a well-defined need for a cryocooler for a biomagnetic SQUID gradiometer. This instrument

would operate in a stationary, well-controlled environment, where

self-generated vibration and magnetic interference could be reduced

by careful mounting and shielding techniques. If an economical and

reliable cryocooler could be produced for this application, it would

make magnetocardiography and pneumomagnetic measurements acceptable

for general clinical use. Experience gained with such systems could

lead naturally to development of portable instruments for geomagnetism and magnetic anomaly detection. Another, less demanding, application for a small cryocooler was suggested by L. Holdeman and C. Chang (p. 243, ref. 1), a simplified portable Josephson voltage standard. Systems such as these could serve as indicators of the viability of small cryocoolers for many other applications, and should not

require large amounts of money for their development.

REFERENCES

1) Applications of Closed-Cycle Cryocoolers to Small Superconducting Devices, edited by J. E. Zimmerman and T. M. Flynn.

National Bureau of Standards, Special Publication 508 (April,

1978) (in press).

2) Closed-Cycle Cryogenic Cooler Technology Applications, Vol. 1,

Tech. Report AFFDL-TR-73-149. Wright-Patterson Air Force Base

(Dec. 1973).

3) 1976 Navy Study on Superconductive Electronics, edited by

A. H. Silver, Office of Naval Research Report No. NR319-110

(Aug. 1976).

4) Superconducting Quantum Interference Devices and their Applications, edited by H. D. Hahlbohm and H. Lubbig. Walter de Gruyter,

Berlin (1977).

5) Superconducting Machines and Devices, Vol. Bl, edited by S. Foner and B. Schwartz, Plenum Press, New York (1974). Primarily

on very large systems.

6) R. J. Dinger, J. R. Davis, and M. Nisenoff, NRL Memorandum

Report 3256. Naval Research Laboratory. (Mar. 1976).

7) V. Vorreiter, private communication. See also ref. 8.

8) C. W. F. Everitt, et al . , To Perform a Gyro Test of General

Relativity. Final Report on NASA Grant 05-020-019. High

Energy Physics Laboratory, Stanford University (July 1977).

67
9) Technology of Liquid Helium, edited by R. H. Kropschot,

B. W. Birmingham and D. B. Mann. NBS Monograph 111. (Oct.

1968) p. 85.

10) T. R. Strobridge. Cryogenic Refrigerators - an Updated Survey.

NBS Tech. Note 655 (1973).

11) G. Prast, Philips Tech. Rev. 26, 1 (1965). J. Kohler, The

Stirling Refrigeration Cycle. Scientific American, 242 , 119

(1965).

12) W. M. Wynn, C. P. Frahm, P. J. Carroll, R. H. Clark, J. Wellhoner,

and J. M. Wynn, IEEE Trans. MAG-11 , 701 (1975).

13) M. McColl, M. F. Millea and A. H. Silver, IEEE Trans. MAG-13 , 221 (1977).

68
International Journal of Refrigeration 2 (1979), pps. 211-213.

Very low-power Stirling cryocoolers using

plastic and composite materials

D. B. Sullivan and J. E. Zimmerman

Cryocables a cycle Stirling de

tres faible puissance utilisant

des materiaux plastiques et

composites

Une e'tude experimental a 6t6 entreprise, concemant

plusieurs conceptions de refoidisseurs cryoggniques pour dispositifs supra-conducteurs ultra-sensibles. Les

appareils a refroidir sont ceux qui utilisent les jonctions Josephson, comme par exemple les magnetometres SQUID, les fralons de voltage, et les transformateurs alternatif-continu. Ces conceptions

possedent toutes la caract&istique essentielle que de

tres bas niveaux d'interfe"rence magnetique, §lectrique

et m6canique sont potentiellement r&alisables.

An experimental investigation of several concepts for very-low-power cryocoolers for operating highly-sensitive superconducting

devices has been undertaken. The devices to be

cooled are those using Josephson junctions,

such as SQUID magnetometers, voltage standards and A/D converters. The common

basic feature of the concepts is that very low

levels of magnetic, electric, and mechanical

interference are potentially realizable.

Nylon and epoxy-glass composites have been

used for cold cylinders, pistons, and displacers.

A multi-stage Stirling machine of these materials

reliably maintains a temperature below 10 K, with a few mW heat load and an input drive power of 50 W. A single-stage machine

maintains a temperature in the neighbourhood of 4K with the hot end at 10K and a drive power of the order of 5W. Specific heats and thermal

conductivities of the materials have been

measured between 5 and 10 K, for purposes of

analysis of regeneration inefficiency.

We have used unconventional materials and operating parameters in a classical type of heat pump to produce a new class of cryocoolers. They operate at low speed, with low input power, use mainly plastic and non-metallic composite materials for the cold

parts, have only milliwatts refrigeration capacity, and

are simple to build. Having such low refrigeration capacity, their initial cool down from room

temperature requires many hours unless special measures are taken to speed up the process. The

ultimate purpose of these cryocoolers is to cool small superconducting devises such as SQUIDS and

Josephson voltage standards, whose cooling

requirements are milliwatts to nanowatts or less.

The cryocoolers to be described here are Stirling machines, but the special features they incorporate could equally well be adapted to Gifford-McMahon

and Viulluemier machines. One of the special features

is the use of a plastic (nylon) and a composite (glass- reinforced epoxy) for the cold parts of the machine and also to provide regenerative heat exchange. In fact the regenerators consist of the annular gaps between the pistons and cylinders with the walls of each providing thermal storage. At an operating

speed of the order of 1 Hz, heat penetration into

The authors are with the Electromagnetic Technology Division, National Bureau of Standards, Boulder. Colorado, 80303. This paper was presented to Commission A1/2 at the XV International Congress of Refrigeration in Venice. Italy. This work was supported by the US Naval Air Systems Command

these materials is the order of 2mm at 4K so that regenerative heat exchange is limited by a thermal

resistance in the material rather than in the working

fluid, as is usually the case for regenerators made of Pb shot or other highly-conducting materials. The

operation of these machines has been analyzed in detail by Radebaugh. 1

One cryocooler described below is a multi-stage machine with nylon displacer and glass-epoxy cylinder which maintains 8.5 K or less with no heat load. The

second machine is a single-stage two-piston

arrangement which maintains 3 to 4 K, with the hot end at 8 to 12 K. This machine operates in the two- phase region. That is, it liquefies helium at the cold end, so the cold-end temperature is the boiling point

of helium at the average operating pressure, and the lowest temperature was achieved with an average

pressure considerably less than atmospheric.

Since both of these machines have been described in the literature,** their essential features will be only

briefly summarized here. Following this we will speculate on their limitations and further developments, particularly on the possibility of combining them into a single unit capable of maintaining temperatures below 4 K in a 300 K

environment.

Four-stage Stirling refrigerator

A schematic of the four-stage machine 2

is shown

69
in Fig. 1. At an operating speed of 1 Hz and an

average pressure of 5 MPa the SQUID attached to the cold end was operated at 8.5 K. A brass gas- thermometer bulb with an internal volume of 0.5 cm 3 was interposed between the SQUID and the bottom

end of the displacer cylinder. The bulb was connected

to a pressure gauge at room temperature through a 0.8 mm inside diameter Cu-Ni tube, and the system was filled with helium gas at a pressure of 0.06 MPa

with the bulb at 8.5 K. The thermal resistance and

heat capacity of the thermometer bulb were intended

to smooth out the amplitude of the temperature

oscillations of the SQUID, relative to the amplitude at the displacer cylinder itself. The amplitude as

indicated by the pressure gauge was about 0.05 K, peak-to-peak.

Although Nb point contacts and SQUIDS can be

operated at 8 to 9 K, their performance and that of several other devices improve at lower temperatures. Probably the main limitation on the cold end

temperature of this machine is the poor regenerator performance. The total regenerator heat capacity of the final stage operating between 8.5 and 20 K is estimated to be 0.1 JK -

' compared to .02 JK _1

for the working fluid. Approximate average

parameters used for this estimate are thermal

conductivity, x = 1.5 mW cm -1 K"\ specific heat, C

= 50 mJ cm'3 K"1

; ra'dius, q = 2.2 mm; length, I =

15 cm, and operating speed, <±> = 2n radians s"\ from which the total heat capacity is~2ng I \f2xClu>. The working fluid heat capacity is simply the specific heat of helium at the peak pressure, estimated to be

0.7 MPa at a cold end temperature of 8.5 K, multiplied by nQ 2 d, where d = 1 mm is the displacer

Plastic piston ond cylinder 36cm displacement

Gas - thermometer [ ^^

gouge

Radiation shields

Gos-thtrmomettr bulb Squid

Fig. 1 4 stage Stirling refrigerator

Fig. 1 Cryorifngirateur a cycle Stirling a 4 itages

stroke. A heat capacity ratio of 5 is too small to provide efficient regenerative heat exchange, and the

ratio becomes rapidly smaller if a lower cold-end temperature is assumed. A simple way to increase the

ratio, using the same materials, would be to provide an additional annular gap or gaps for additional regenerator area, as was done in the single-stage machine described in the following paragraph.

Single stage twin piston Stirling refrigerator Another way to improve the lower stage performance

is to reduce the helium pressure so as to enhance the

ratio of the regenerator-to-helium heat capacity. This

will also reduce the available cooling power, but for small heat loads the temperature achieved can be

lower. This approach did not work for the multistage machine because the larger refrigeration power at higher pressure was essential for the upper stages. However, a single-stage low-pressure machine could be operated as a separate additional stage at the bottom of the higher pressure multistage device. To

test this idea, we built and tested a single-stage twin- piston Stirling machine 4 which produces several mW

of refrigeration in the 3 to 4 K range with upper end

temperatures of 8 to 12 K.

Fig. 2 is a cross sectional drawing of the device. The

two pistons operate concentrically inside a fixed cylinder and the movement of helium gas between

the bottom end and the room temperature housing is limited only by the small radial clearances between

the three concentric pieces. All three pieces were made from commercially available spun-glass epoxy CG— 10'). At room temperature the viscosity of helium is large and thus little gas moves between the upper housing and the cold parts during each cycle of the machine (the speed is 1 Hz). On the other hand,

the reduced viscosity at low temperature results in large flow and thus little pressure drop between the compression and expansion spaces. As shown in the

figure, this machine has two regenerator gaps in parallel to further enhance the heat-capacity ratio.

An electric heater was mounted on the hot end flange which was in turn connected through a weak thermal

link to a liquid helium bath. With this arrangement the hot end could be maintained in the temperature range

of 8 to 12 K. A second heater was mounted on the

cold end to determine refrigeration capacity. Carbon

resistance thermometers were used to monitor the temperatures at the cold and hot ends as well as at several intermediate points. Changes in the helium

boil-off rate were used to determine the magnitude of the heat rejected at the hot end.

For most of our experiments the temperature of the

cold end was limited by entry into the two-phase

region. The presence of liquid (of order 0.01 cm 3

) was confirmed by observation of the warming rate of the cold end after the refrigerator was stopped. The

cold end temperature would typically remain at that of the liquid phase for several minutes and then rise quickly after the last of the liquid evaporated. With

this confirmation of two-phase operation, it was clear that the operating temperature was essentially the

70
equilibrium vapour pressure of the helium gas which was measured with the gauge shown in Fig. 2. A

temperature of 3.1 K was achieved at an average

pressure of 0.027 MPa (200 mm Hg) with the hot end

at 8 6 K With the hot end at 12 K and a pressure of 0.08 MPa (600 mm Hg), the cold end reached a temperature of 4 K. The heat ejected at the upper end

(for this latter experiment) was about 50 mW and the

refrigeration power at 4 K was 4 mW.

We have measured the heat capacity and thermal

conductivity (these will be reported elsewhere) of the spun-glass epoxy regenerator material and it is instructive to see how these influence the regenerator performance. For the following example we take the operation from 12 to 4 K as representative of the

cooler operation. The diffusivities, D = xlC, at 4 and

12 K, respectively, are 0.36 and 0.063 em's2

. For 1 Hz

operation the thermal penetration depth is 3.3 mm at the cold end and 1.2 mm at the hot end. The bottom

end value indicates that a substantial portion of the

available material is indeed used for heat storage at the lower end. The heat capacity of the entire regenerator surface (including the gaps inside and

outside the lower stationary insert) is approximately

0.17 JK" 1. The average value of gas heat capacity at 0.08 MPa (600 mm Hg), is 0.0087 JK"' so that the

ratio of regenerator capacity to heat capacity of the gas is~20. For comparison, the ratio would be only about 2 if the operating pressure were near 1 MPa (as

in the multistage device). The need for reduced

operating pressure is thus obvious.

P. Roubeau (private communication) has pointed out

that gravity separation of the phases may play an

important role in the operation of this machine. Being

relatively incompressible, the liquid which collects in the bottom of the expansion space no longer takes

part directly in the refrigeration cycle, but by filling the dead space and providing thermal ballast it may

significantly increase the efficiency of the process. This is not completely certain, since the non-ideal

characteristics of the fluid could offset this gain 1

. Obviously this action would depend on the attitude of the machine relative to gravity. Future developments

For many purposes, an improvement of the multistage

refrigerator (as discussed earlier) to achieve a temperature of 6 or 7 K will be adequate. However, to reach 3 or 4 K will require a combination of the two

machines. The overall size and power requirement of the combination will be smaller if the low pressure stage is reduced in size and if it operates over a narrower temperature difference. If the diameter were

reduced by a factor of two, then the volumes and

heat ejected at the warm end would be four times

smaller. The reduced cooling capacity could be compensated in part by an improvement in the performance of the multistage machine. As discussed

earlier, this would require an increase in the overall heat capacity of the regenerator material, particularly

in the bottom stage. It is thus conceivable that the combination might be comparable in size and power

requirement to the multistage machine described above.

ORingseol ^v En<j „„,, „, (\ .y crcnkshafl Crankshofti ^—

Connecting rods gxt 4 Helium ' ,l1

Pressure gouge

Clearances ~ 30^m

Lower piston -*- 6mm 00

Upper piston (annular) ~i2mmOD

Outer Cylinder I6mm00

Hot end flange 8K-I2K

Compression space

I cm stroke

Fleiible joints

Stationary insert

Regenerator gaps -^ 30^m

Expansion space Icm stroke

-Cold end

Fig. 2 Single stage twin piston Stirling refrigerator. The flexible joints permit self-centring of a centre piston so as to reduce

fnctional heating

Fig. 2 Cryoretrigerateur a cycle Stirling a double piston a un etage Les joints flexibles permettem un auto-centrage du piston central

afin de riduire I'ichaultement par frottement

With regard to the geometry of the combination, the two could be operated side by side with the hot

flange of the single-stage coooler connected to the

cold end of the other machine. The two would

obviously operate from a single drive motor. However, it might be advantageous to consider a concentric configuration with the single-stage machine within the multistage one. Besides simpler machine work the heat conduction in the single-stage device would naturally be intercepted by higher temperature stages in the multistage machine. Also, the regenerator area of the multistage machine

(particularly at the lower end) would be increased because of the larger diameter necessary to accommodate the coaxial single-stage machine. References

1 Radebaugh, R. Analysis of regenerator inefficiency for Stirling- cycle refrigerators with plastic displacers. Proceedings of the XV

International World Congress of Refrigeration (1979) 2 Zimmerman, J. E., Radebaugh, R. Operation of SQUID in a very low-power cryocooler, NBS Special Publication 508 11978) 3 Zimmerman, J.E.. Radebaugh, R., Siegwarth, J.D. Possible cryocoolers for SQUID magnetometers. Superconducting Quantum

Interference Devices and Their Applications (H. D Hahlbohm and

H. Hubbig, eds) Berlin, Walter de Gruyter H977) 287

4 Zimmerman, J. E., Sullivan, D. B., A milliwatt Stirling cryocooler for temperatures below 4 K Cryogenics (March 1979)

71
Proceedings of the SQUID Geophysics Workshop, Los Alamos, NM (June 1980)

CRYOCOOLERS FOR GEOPHYSICAL SQUID MAGNETOMETERS1

"

James E. Zimmerman

National Bureau of Standards

Electromagnetic Technology D1vts1on

Boulder, CO 80303

ABSTRACT

No alternatives to Hquid-helium cryostats for SQUID geomagnetic measurements are presently available, but micro-miniature Joule-Thomson and low- power non-magnetic Stirling cryocoolers are being

developed for this and similar purposes. With

increasing interest and experimental work on the subject during the past year or two, it 1s likely

that demonstrations of feasibility will occur in the moderately near future, and perhaps even a suitable commercial cryocooler 1n the next few years.

TEXT

To design a small closed-cycle cryocooler for use with geophysical SQUID magnetometers will be much more difficult than for a gradlometer, since

the latter 1s Insensitive to motion (1n particular,

rotation) in the earth's field caused by the cryocooler. Clearly, the problem 1s not Impossible.

That is, a system might consist of a small helium

liquefler, probably operated Intermittently, to maintain liquid helium 1n a magnetometer dewar,

the vapor being collected for recycling. A system

of this type has been operated 1n the laboratory

(Vincent, 1978). While this type of system clearly

is not what most of us have 1n mind when thinking

of a cryocooler for a SQUID magnetometer, perhaps

it should not be rejected out of hand. A fraction

of a liter volume 1s sufficient for a 3-ax1s

magnetometer, and by efficient design a dewar of

this size could operate at least 24 hours on one

filling. A Hquefier operating 10 hours per day

should therefore have a capacity of the order of

0.05 L/hr. This would require 6 W of mechanical

input power for a perfect liquefler, or 120 W 1f one assumes an efficiency of 5%. A gas holder

for the gas equivalent of 1/2 liter of liquid

would be somewhat bulky, about 1/2 m^ at 2000 m above sea level

.

Wtth regard to efficiency, a statistical survey (Strobrtdge, 1974) several years ago sug- gested that efficiency of helium Hquefiers and refrigerators drops dramatically at refrigeration

capacities less than a watt or so at 4K. Perhaps

this trend should not be taken too seriously,

however, wtthout more specific studies on the ltmttattons of small, low-power cryocoolers. A SQUID cryocoeler Built 1n NBS-Boulder (see below)

has an efficiency of about ML (8 mW refrigeration

at 3 K with an estimated electrical power Input

of 30 W), which 1s still comparable to the ef- fldences of cryocoolers 10 to 100 times larger,

at the lower end of the statistical survey.

The total heat leak, and the size of the system, mtght be reduced significantly by In- corporating the ltquefler Into the magnetometer

dewar Itself. Also, the dewar capacity can be reduced 1f the liquefler 1s operated a few

minutes out of each hour rather than a few hours out of each day, as suggested above. Although

no systems of this type have been built, 1t 1s probably the one most commonly envisioned

as a compromtse between the Inconvenience of

the conventional cryostat requiring periodic

transfer and filling from a (usually distant and expensive) source of liquid helium, and the difficulty of designing a continuously running but non- Interfering cryocooler. It 1s obviously limited

In application to short-term intermittent operations, and does not easily lend itself to operation 1n arbitrary orientations.

The tdeal system, from the point of view of

efficiency, portability, and flexibility is a continuously runntng cryocooler. Currently, two cryocooler types with some potential of being

adaptable to geophysical SQUID magnetometers are

being developed.

Contribution of the U. to copyright.

S. Government, not subject

72
One of these is the micro-miniature JouleThomson cooler proposed by Little and subsequently

investigated experimentally using silicon chip

microcircuit technology to form the channels for the counter flow heat exchanger and the JouleThomson throttle (Little, 1978 a and b). The high

thermal-conductivity of silicon makes It impractical

for very low-temperature refrigeration, and recently

attention has been directed toward adapting the idea and techniques to low-conductivity materials such

as glass. A practical cryocooler based on Little's

Idea has not yet been reported, so we will speculate

on some of its potential advantages and disadvantages.

First, and maybe foremost, there are no solid moving

parts at low temperature, and it was pointed out by

Little that fluid flow 1n the narrow (<v 10 ym) channels of a microminiature Joule-Thomson cooler should

be laminar. Consequently, there should be no high

or low-frequency noise generated by this part of the system - a matter of crucial importance for SQUID

magnetometers. Second, the essential low-temperature

elements of a Joule-Thomson cooler, the counter flow

heat exchanger and the throttle section, can be made of very narrow channels which might be built

into the SQUID itself and its electrical leads, so that the system cools precisely the part that needs

to be cooled and nothing else. This 1s the essence of Little's idea, to develop a flexible technique

for matching the cryocooler to the requirement of

a specific device. Third, owing to the inherent

low heat capacity of the system, cool-down time

could be of the order of 1 minute. This would be

ideal for applications like geothermal prospecting

where measurements are Intermittent and shortterm.

On the other hand, the narrow channels of the Joule-Thomson system may be easily clogged with

Impurities (frozen air, grease, dust particles, etc.)

which may be difficult to remove. Second, at least

three separa.te stages (N-, H

2

, and He, for example)

are required to reach the operating temperature of

presently available SQUIDs, and rather high pressures

are required for the higher temperature stages. For many applications, gas from high-pressure cylinders

could be used, and Little has estimated 1n a par- ticular example an operating time of 400 hours from

a standard cylinder of gas. Third, a closed system

for continuous long-term operation would require

three compressors. These could be remotely located

to avoid interference, but the relatively high pressures and stringent limits on solid or condensable

impurities would tend to make the system rather

complicated and expensive. Fourth, the Joule-Thomson

effect 1s thermodynamlcally Inefficient. However,

the inefficiency might be more than compensated for

by optimally matching the cryocooler to the SQUID

as suggested above.

A quite different SQUID cryocooler 1s being

tested 1n the NBS-Boulder Laboratories (Zimmerman,

et al., 1980, 1978a, 1978b, 1976, Sullivan and

Zimmerman, 1979). This 1s a standard Stirling

cryocooler , with some unique features to adapt 1t for use with SQUIDs, operating in the neighborhood

of 7 K. In its present version, shown schematically

1n Fig. 1, a solid nylon stepped dlsplacer moves up and down in a matching epoxy-glass composite cylinder.

This type of cryocooler is simple, easy to build,

easily portable, operates on rather low Input power,

and 1s potentially capable of long-term continuous

unattended operation. An earlier version operated

a total of more than 7000 hours (not continuously)

without observable wear or degradation of performance.

The problem with this machine, which over- shadows all others, 1s that It directly and In- directly generates magnetic Interference. Tests made with a transverse SQUID magnetometer, mounted

on the cold end of an earlier version of the machine,

showed that a peak-to-peak signal of about 0.5 nT was generated by rotation of the SQUID 1n the earth's

field due to flexing of the dlsplacer cylinder at the cryocooler operating frequency of 1 Hz. There were also unidentified higher frequency components

with an amplitude of the order of 10 pT. With the present gradiometer-cryocooler, a rather clean 1 Hz interfering signal was measured with an amplitude

corresponding to a peak-to-peak field of about

0.3 pT 1n the gradlometer loop nearest the cold

end of the dlsplacer cylinder. Since the gradio- meter may be assumed to be Insensitive to rotation

1n the earth's field, this interference 1s probably

caused by the dlamagnetic moment of the moving nylon

dlsplacer In the earth's field, or by the changing

paramagnetic moment of the glass-epoxy cylinder

walls due to local temperature variation during the refrigeration cycle. We have determined that the susceptibility of nylon 1s about the same as that of quartz Cabout 10" 5

).

Although the Interference components identified

above are rather large, it should be appreciated

that "clean" signals at the cryocooler frequency

and multiples thereof can be largely eliminated by any of various methods of subtracting the interference

from the magnetometer output. By this and other

methods of interference rejection, 1t appears that the system can be made practical for some of the less-demanding applications In geophysics, perhaps

at the ptcotesla level of sensitivity.

It would be difficult to guess when practical

cryocoolers wtll be generally available for geo- magnetic measurements. In the past two or three years several other laboratories have started

experiments on both microminiature Joule-Thomson

coolers and on plastic Stirling cryocoolers similar

to the above, so 1t Is conceivable that something

may be commercially available within a few years.

*The Sttrling machtne, whtch 1s probably the stmplest and most efficient small heat engine for large temperature differences, was Invented 1n 1816 and used to pump water out of a mine. The first Stirling refrigerator was butlt about 1860.

It was htghly successful CWalker, 1973).

73
Work on the Stirling cryocoolers 1s being done

1n collaboration with D. B. Sullivan, and 1s sup- ported 1n part by the Office of Naval Research and

the Naval Coastal Systems Center.

REFERENCES

Little, W. A., 1978, Design and Construction of

Hlcromlnlature Cryogenic Refrigerators, Future Trends 1n Superconductive Electronics , edited by

B. S. Deaver, Jr., et al., AIP Conference

Proceedings No. 44, p. 421.

Little, W. A., 1978, Scaling of Miniature Cryocoolers

to Microminiature Size, NBS Special Publication

508, April 1978, p. 75.

Sullivan, D. B. and Zimmerman, J. E., 1979, Very low- power Stirling Cryocoolers using plastic and

composite materials, International Journal of

Refrigeration, v. 2. p. 211-213.

Strobrldge, T. R., 1974, NBS Tech. Note 655.

Vincent, D. A., 1978, Closed-Cycle Refrigerator for

a Superconducting Susceptometer, NBS Special

Publication 508, April 1978, p. 131.

Walker, G., 1973, Stirling-Cycle Machines . Oxford

Press.

Zimmerman, J. E. and Radebaugh, R., 1978, Operation

of a SQUID 1n a Very Low-Power Cryocooler, NBS Special Publication 508, April 1978, p. 59.

Zimmerman, J. E., et al., 1976, Possible Cryocoolers

for SQUID Magnetometers, Superconducting Quantum

Interference Devices and their Applications , editors, H. D. Hahlbohm and H. LUbblg, Walter

de Gruyter.

Zimmerman, J. E., 1978, Crycoolers for Superconductive Electronics. Future Trends 1n Superconductive Electronics , edited by D. S. Deaver,

Jr., et al., AIP Conference Proceedings No. 44,

p. 287.

Zimmerman, J. E. 1980, Cryogenics for SQUIDs,

Presented at the Second International Conference

on SQUIDs, Berlin, May 6-9, 1980. To be published.

Preprints available.

7 mm

Stroke

Common Cmfcshsfl •nd Motor Drivt -A

""

Solid Nylon Dttpiacar

(5 Sactions)

Epory—Giaa Cylinder

Raganarator Gap

( Exaggerated)

Expansion Spam

Vacuum Spaca

(Radiation Shialdi

(90 em3 Ditpj,;

6 em (Approx.)

Gradiomatar AaMmbly

and Squid

Figure 1. -Simplified schematic cross- sectton of SQUID cryocooler, approximately to scale. Note different horizontal and vertical scales. Non- magnetic crankshaft and crosshead and magnetically

shielded 50 W drive motor not shown. Total weight

about 20 kg. (See Zimmerman, 1980 for detailed

specifications and performance).

74
"SQUID 80: Superconducting Quantum Devices and their Applications," (H. D. Hahlbohm and H. Liibbig, eds.), Berlin, Walter de Gruyter (1980), pps. 423-443.

CRYOGENICS FOR SQUIDs

James E. Zimmerman

Electromagnetic Technology Division

National Bureau of Standards

Boulder, Colorado 80303 U. S. A.

Abstract

Liquid-helium cryostats continue to be the principal

mechanism for cooling SQUIDs and SQUID instruments, but

there is increasing interest in self-contained closedcycle cryocoolers as an alternative. Mechanical gas

refrigerators, in various configurations, and the JouleThomson effect seem likely to be the most practical

mechanisms, but magnetic, electric, and other processes

may be useful if suitable materials can be developed.

While not much work on SQUID cryocoolers as such has

been done yet, a large amount of work on higher temperature refrigerators for infrared sensors and for use in

spacecraft provides a sophisticated technical base upon

which to build. A low-power five-stage Stirling cryocooler

with a single-component SQUID gradiometer has begun

operation in the author's laboratory. Although this

machine requires only 20 W of mechanical power input

to maintain a temperature of 7 K, a large reduction of

input power is theoretically possible.

Introduction

Liquid-helium cryostats continue to be universally used to cool

SQUIDs and other superconducting devices. During the past year

or two, however, there has been increasing interest, both in

government funding agencies and in private industry, for the

75
development of small cryocoolers for SQUID magnetometers and

similar instruments, and of somewhat larger machines for

superconducting computers (1), (2).

For many years there has been extensive development of small

refrigerators for infrared detectors at higher temperatures,

above 20 K, (3) and for masers (2). Much of this technology

and experience will be directly useful in the development of

cryocoolers for SQUIDs. Unfortunately, much of the literature

is in the form of government reports, not readily available,

even though unclassified.

A short but imaginative and comprehensive review of refrigeration principles and practice has been written by Radebaugh (4)

.

A very comprehensive treatise on mechanical gas refrigerators

called "Cryocoolers" is in preparation by G. Walker for

publication within a year by Plenum Press. It contains more

then 1000 references. The various proceedings of the Intersociety Energy Conversion Engineering Conference are a rich source of information and ideas on all kinds of heat engines.

A recent review on cryogenics for spacecraft, including cryocoolers, solid and liquid cryogens, and passive radiation

cooling, has been written by Sherman (5) . The extreme

reliability and other special requirements for cryogenics

in space has resulted in improvements in bearing and seal

design, materials, portability, and efficiency, which will

be useful in the design of cryocoolers for SQUIDs.

A. MACHINES AND PRINCIPLES

Most work in the U. S. A. and elsewhere on small cryocoolers,

below 20 K, and on higher temperature machines has been on Stirling, Vuilleumier, and Gif ford-McMahon machines, shown

schematically in Fig. 1. All of these use displacers and

regenerators. The regenerator may be built into the displacer,

76
s»

*t~ Piston No Work

J*x

jr~ v q2 *w

Displacer

and Regenerative Heat Exchanger

Cold

ssz.

©

Cam

JB

JCompressor

7

3lic%^.

Cold

£n.

•

a«

Cold

Compressor

/"'

Hot

Counterflow Heat Exchanger

I?

Throttle Valve Cold

Fig. 1. Schematic representation of Stirling (upper left) , Vuilleumier (upper right) , Gif ford-McMahon (lower left) and

Joule-Thomson refrigeration systems.

as shown, or into the walls of the surrounding cylinder, or

as a separate unit. In the simplest machines (see, for

example, Appendix A) , regenerative heat exchange takes place

between the surfaces of the displacer and of the cylinder

and the working fluid in the intervening gap. In the Stirling

machine the displacer and piston are driven in synchronism,

with a phase difference of the order of tt/2, and likewise

for the two displacers of the Vuilleumier machine. Both the

Stirling and the Vuilleumier are reversible in the sense

that they can pump heat either direction, and a Stirling

machine can be used with a drive motor as a refrigerator or

as an engine delivering work to an external load. In a

77
Gifford-McMahon machine the compressor operates independently

of the displacer, and the pressure in the displacer cylinder

is raised and lowered through valves appropriately synchronized

with the motion of the displacer. This type of machine is

not simply reversible, and is probably less efficient than

the reversible types. It has a significant advantage, for

SQUID users, that the potentially noisy compressor can be

remotely located. All of these types of machines have

achieved temperatures in the neighborhood of 6 to 7 K, using

multi-stage or stepped displacers. For temperatures down to

4 K, a Joule-Thomson final stage can be added. An alternative

to a Joule-Thomson stage has been proposed (6).

There are many other recognized cooling machines (Solvay,

Claude, etc.), mainly using helium gas, or sometimes hydrogen,

as the working fluid. Then there is a spectrum of mechanisms

using magnetic, or perhaps electric, polarizable material as the working fluid, and machines operating in the neighborhood

of critical temperature of any of various fluids. "Evaporative"

coolers such as home refrigerators and helium liquifiers

using the Joule-Thompson effect in the final stage are examples of the latter.

Magnetic cooling techniques (an extension of classical methods

of achieving temperatures below 1 K) have been under study

for several years both at higher temperatures, and, more

recently, in the range of 2 to 20 K (7, 8, 9). These might

be difficult to use with SQUID magnetometers, however,

because of interference. Paraelectric cooling of SQUIDs is an interesting possibility, but suitable materials have not

yet been found (10)

.

Refrigeration is a difficult subject to deal with summarily.

Thermodynamics provide a set of simple and elegant general

principles on the minimal conditions necessary to achieve

refrigeration and on the maximum refrigeration rate that can

78
be achieved under specific conditions, but offers no clue as

to a possible mechanism to achieve that rate. For example, a source of heat Q at an elevated temperature T. , such as the sun or a piece of coal, can be used to remove heat, that is,

to refrigerate, at a rate Q

2 from a region at a low temperature

T

2

, where both Q and Q are ultimately rejected to the

environment at temperature T . The quantitative relation

for the maximum possible refrigeration rate is

T T — T

Q 2 = Qi t7 T^rf • *U

1 o 2

An example of a refrigerating mechanism is a thermoelectric

generator connected to a Peltier cooler - two thermocouples

in parallel with the connecting wires maintained at ambient

temperature. With conduction electrons as the only moving

parts, such a mechanism should be highly reliable. Suitable

materials for practical cryocoolers using this mechanism do

not seem to have been discovered, however. A more practical

example is the Vuilleumier machine (11). It has moving

parts, but no work is done on or by them, so mechanical

forces are small. Such machines are completely self-contained in the sense that they can be operated with "primary"

energy sources such as the sun or petroleum-derived fuels.

Thus, they may be attractive for use in space and other

remote locations.

When a source of mechanical work, such as an electric motor,

is available, then the maximum possible refrigeration Q 2 is

related to the mechanical power W by

6

2 = w T-^hr (2)

79
where T

2 and T are as defined above. Most of the betterknown refrigerating machines, Stirling, Gif ford-McMahon, and

so on, are of this class. They are of interest to the

majority of those who are located near "secondary" energy

sources such as electric power lines.

Two commonly used idealizations of thermodynamic cycles

achieved by the above machines are the Stirling cycle and

the Ericsson cycle. The former consists of an isothermal

compression in the warm (room- temperature) end of the machine, a constant volume displacement of gas from the warm

end to the cold end through the regenerator, an isothermal

expansion, and a constant volume displacement back to the warm end. The Ericsson cycle consists of the same two

isothermal processes, but the processes of displacement are done at constant pressure rather than constant volume. A

Stirling machine can approximate the Stirling cycle if it is

operated slowly with a phase difference of tt/2 between

piston and displacer motions. It can equally well approximate

the Ericsson cycle, if the phase difference is shifted so the compression process partly overlaps the downward displacement (from warm to cold ends) of the working fluid and

the expansion process overlaps the upward displacement.

E. Interference With SQUID Operation

Cryocooler parameters of interest to users of SQUIDs are

cost, reliability, ease of use and operation, cooldown time,

refrigeration rate, power input, size and weight, temperature

stability. Above all, users of SQUIDs will want to know

whether cryocoolers can be developed which will not interfere

significantly with their operation. SQUID magnetometers are being used in geophysical applications where very low frequency

fields, below 1 Hz are measured to 1 pT or better. Gradiometers

used for geomagnetic anomaly detection at frequencies below

1 Hz and for biomagnetism between one and 100 Hz require

80
noise and interference levels at the gradiometer pickup

loops of 0.1 to 0.01 pT or less.

Operation of a SQUID in a cryocooler has been reported once previously. An interference level of 0.5 nT at the 1 Hz

operating speed was measured (12) . It was shown that most

of this was due to motion of the cold end of the cryocooler,

on which the SQUID was mounted, in the earth's field.

Although the problem of reducing the interference level by

several orders of magnitude is formidable, there are several

ways of dealing with it, none of which have been exploited

up to now. First of all, some careful attention to the

cryocooler structure might result in at least an order of

magnitude reduction in motion of the cold end. For SQUID

magnetometers used in low dc fields, as in interplanetary

space or in the shielded rooms of Massachusetts Institute

of Technology, Physicalisch-Technische Bundesanstalt, and

Helsinki Technical University, motional interference would

be reduced by a factor of 1000.

For a well-balanced gradiometer, motional interference in a uniform field would be reduced by up to 10 , and it should

be possible to reduce direct magnetic pickup from the cryocooler by factors of 10 to 100 relative to that picked up

by a magnetometer.

Since most interference, by whatever mechanism, from a cryocooler should be at the cryocooler frequency and its harmonics,

it can in principle be largely eliminated from the SQUID

output by a "commutative" filter synchronized with the

cryocooler frequency. These filters are used very effectively

in several laboratories to eliminate power-line interference

from the output of biomagnetic SQUID gradiometers. If the

magnitude of the interfering signal is essentially constant,

the filter bandwidth can be made extremely narrow, so that

81
virtually no information is lost from a typical signal.

Since commutative filters may be unfamiliar to many, a primitive version of one is given here (Fig. 2). The rotating

switch or commutator is rotated in synchronism with the

interfering signal at oj that is, at the cryocooler speed.

In modern practice, of course, the switching is done with

solid-state integrated circuits (13) . The rejection bandwidth

of the filters at a and its harmonics is Aco = 1/RC

.

o

Differential Amplifier

-VWWR

/T

Out

o „rf o

O (jj I

etc.

It

Fig. 2. Schematic of primitive version of a commutative filter

for eliminating interference at co and its harmonics.

C. Reliability

One of the most commonly expressed concern about cryocoolers

is reliability. In private discussions, knowledgeable

individuals have often stated that cryocoolers are not

reliable. Some of these statements seem to be based on military-sponsored projects which have fallen short of the

goal of producing cryocoolers for several years of continuous

unattended operation. Most of us do not require such extreme

82
performance, and in any case, users of cryocoolers for

ground-based maser preamplifiers have reported outstanding re- liability (2) . Users of SQUID magnetometers may know that the

reliability of liquid-helium cryostats leaves much to be desired, especially outside the laboratory. In 1973, for example, Wallace Campbell and the author measured the geomagnetic

field variations during the solar eclipse of 1973 in southern

Chad, near the geographic center of Africa (14) . More

recently, some measurements were made on electric currents

in the Alaska pipeline induced by ionospheric currents (15)

.

In each case, only one of two storage containers of liquidhelium arrived at the specified destination. These and

similar experiences have demonstrated the need for selfcontained cryocoolers of modest reliability.

D. NBS Cryocooler Development

The first paper on cryocoolers for SQUIDs was presented 3 1/2

years ago here at this conference (16) . Since then, we have

demonstrated that a simple, potentially low-cost cryocooler

can be built which will maintain temperatures in the range

of 7 or 8 K in which Nb SQUIDs operate very well (17) , (also see Appendix) . We have also built an even simpler, singlestage unit which achieved temperatures in the 3 to 4 K range

with the hot end in the neighborhood of 10 K (6, 17).

Refrigeration capacities of all our machines have been on

the order of a few mW per Kelvin temperature rise of the

cold end, with correspondingly larger cooling capacities at

the higher-temperature stages. Emphasis in all our work has

been on low power and low interference levels, for compatibility with SQUID magnetometers and small gradiometers . One

result of this emphasis has been machines maintaining 7 or 8 K with driving power levels very much lower than those

required for previous machines in this temperature range.

Our machines require 50 to 100 W electrical power, using

83
inefficient ac drive motors. With reasonably efficient dc

motors and other refinements, it is probable that temperatures

below 10 K could be achieved with 10 to 20 W of electrical

power

.

While it has been rewarding to demonstrate that superconducting and even liquid-helium temperatures can be reached

with low-cost materials and simple mechanisms requiring no great skill to assemble, the discussion which follows indicates

that large improvements in efficiency may still be possible.

E. Efficiency

Efficiency is not usually of prime concern in the design of

small cryocoolers. Still, it might be emphasized once more

that the inherent refrigeration requirements for SQUIDs are smaller by orders of magnitude than that provided by most

"small" or "miniature" cryocoolers.

Even though the cost of energy to run a cryocooler may not be

significant (even these days) , increasing the efficiency so

as to reduce the size and power requirements of a SQUID cryocooler has potential benefits in terms of reduced capital

cost, portability, and reduction of interference. In the

following section are presented some numerical comparisons

between liquid-helium cryostats, the NBS small cryocoolers,

and the inherent refrigeration requirements for SQUIDs.

F. Liquid-Helium Cryostats, the NBS Cryocooler, and

Inherent SQUID Requirements

A "super-insulated" helium dewar of 1 L capacity with vaporcooled radiation shields can have an evaporation rate of the

-3 3 order of 5.8 x 10 cm /s (.1/2 L/day) . If we were to con- tinuously replenish this liquid helium by the use of an ideal helium liquefier, the amount of mechanical power W

required would be

84
300 K

- T T - T

W = m !<•'> / C -Z-— dT + mL —

4 K

(3)

where m is the mass evaporation rate, C is the specific

P - heat at constant pressure, T

Q

= 300 K, and L is the latent

heat of evaporation. For 1/2 L/day, the result is approximately

W = 5 W. By comparison, the present and previous NBS cryocoolers have required 10 to 20 W mechanical power to maintain

volumes considerably smaller than 1 L at temperatures of 7 or 8 K. We might conclude that these machines are reasonably

efficient, but that there is room for significant improvement.

There are at least four sources of inefficiency in the

cryocoolers not present in a liquid-helium cryostat, namely,

shuttle heat loss (due to cyclic mis-matching the temperature

distributions along the displacer and cylinder surfaces)

,

regenerator inefficiency (due to imperfect heat exchange

between working fluid and regenerator) , non-isothermal com- pression and expansion (also due to imperfect heat exchange)

,

and mechanical friction.

On the other hand, liquid helium is a very non-ideal cryogen,

because the refrigeration capacity is nearly uniformly

distributed in temperature. For SQUID and other small

device applications, the essential refrigeration requirements

are heavily weighted toward the room- temperature end of the

structure. An outer radiation shield at 150 K, for example,

should require more than 16 times as much refrigeration as one at 75 K receiving radiation from the outer shield at 150

K. Similarly, thermal conductivity of most alloys and

plastics increases with temperature. To demonstrate the

effect of a more efficient distribution of cooling capacity,

let us assume that the absorption of enthalpy by 1/2 L/day

of evaporating liquid helium is proportional to the temperature

of the structure (including electrical leads) at the point

85
where absorption takes place. The total enthalpy change,

from liquid helium to gas at 300 K, is 6300 J/mol., which

gives a total enthalpy absorption rate H = 1.14 W at an evaporation rate of 1/2 L/day. The mechanical power W

required for an ideal refrigerator to provide the same distribution of refrigeration is

300 K .

. T - T

W » / Q -2= dT, (4) /4 K

where Q = AT and A is chosen to satisfy the requirement

300 K

H = / Q dT = 1.14 W. The result is W = 1.13 W, more

"^4K than four times smaller than the mechanical power required

to produce 1/2 L/day of liquid helium.

Our cryocoolers compare rather unfavorably with this more realistic idealization, especially when the relatively small

cold volume of the cryocoolers is considered.

Most cryocoolers have been or are being designed for some specified "net" or "available" refrigeration at one or two

specified temperatures, such as 40 W at 60 K and 1.5 W at 12

K.

For a SQUID cryocooler, it may be more economical not to

specify any available refrigeration, but merely to require

that the cryocooler maintain itself and the SQUID, with its

bias and signal lead wires, at the operating temperature.

As a final exercise on theoretical refrigeration requirements,

let us consider the ideal mechanical power W required to

handle the heat conducted down the wires, assuming an electrical

86
resistivity proportional to temperature, as might be the case for somewhat impure copper, with a room-temperature

resistance for all the wires in parallel of 0.1 ohm. This

corresponds, by the Wiedemann-Franz law, to a temperature-

-5 independent thermal conductance K = 7.3 x 10 W/K. Consider

two refrigeration systems which might be used to maintain

opposite ends of the wires at 4 K and 300 K: (1) the worst

possible, where all of the heat is pumped from k K to 300 K,

and (2) the best possible, where the refrigeration is optimally

distributed along the length of the wires. In the first

case

T - T(0)

W = - T(Q) K (T

q

- T(0)) = 1.6 W, (5)

where T(0) = 4 K and T = 300 K, and the temperature distribution along the wires is linear. In the second case, R. L. Kautz

and D. B. Sullivan have determined by variational calculus

that the optimum temperature distribution is exponential.

T(z)

T(o)

i z T /T(0)

o (6)

where z is the relative distance along the wires measured

from the cold end. The ideal mechanical power is

(

T

°

' T

k

d

' T

dz + T

°

" T

K ) — K ~2- dz + ip— K

o dz

dT

dz

z = o (7)

312 mW + 93 mW = 405 mW.

87
The difference between the two cases is dramatic, especially

considering that the operating resistance would be much

lower in the second case, 0.023 ohm, than in the first case,

0.051 ohm, owing to the lower temperature distribution along

the wires. The numerical values are derived from

R = I ^rr- :rr dz, (8)

/dR dT , dT~ 3£ dz '

dR/dT being determined by the assumptions of a room- temperature

value R = 0.1 ohm, and resistivity proportional to temperature.

A continuous distribution of refrigeration along a linear

structure can be achieved in principle through the use of an elongated, variably tapered expansion space. An experiment

with a long, conically tapered expansion space has been done

by du Pre and Daniels (18)

.

The above discussion suggests a class of problems in optimal

refrigeration designs, where conduction, radiation, and

other heat inputs to the cold device, the wires, and to the

refrigerator structure itself are dealt with as in the above

example. Such problems are very complex, but the approach

might lead to SQUID-cryocooler systems quite different from

those contemplated so far. TJie work of W. Little might be

cited as an inspiration in this particular context.

G. Microminiature Joule-Thomson Cooling

Little has discussed the problem of scaling down the size of

several different refrigeration systems by a factor of 100 or more (19) . In particular, he has studied analytically and

experimentally the possibility of making "microminiature"

88
Joule-Thomson refrigerators by microcircuit techniques on silicon or other suitable substrates, to cool a thin-film

device which might be part of a larger, room- temperature

microcircuit on the same substrate (20) . Silicon-chip

technology has not been found suitable for cryogenic temperatures because of the large thermal conductance of the

silicon, and present work is directed toward adapting the

idea and techniques to other more suitable materials.

The fluid-flow channels in a microminiature Joule-Thomson

cryocooler are so narrow, less than 0.1 mm, that the counterflow heat exchanger might, in principle, be made to resemble

a thin wire, with lower heat conductance than that assumed

for the electrical wires in the numerical example above.

Thus, the mechanical power necessary to handle the intrinsic

heat leak of a Joule-Thomson cryocooler plus the refrigeration requirements of a SQUID together may be very small, in

spite of the inherent inefficiency of the Joule-Thomson

process

.

It was also pointed out by Little that the fluid flow in the

channels of a microminiature Joule-Thomson system should be

laminar and so should generate no microphonic or other

interference, a matter of some interest to SQUID users.

H. Recent Developments in Mechanical Gas Refrigerators

The traditional drive mechanism for mechanical gas refrigerators is a crankshaft and connecting rods. However, some very interesting developments currently in progress are on

the use of linear reciprocating magnetic drive motors directly

connected to pistons and displacers, magnetic or gas bearings

and suspensions, clearance seals, and mechanical resonance to

reduce drive forces (5, 21, 22, 23, 24, 25). The term clear- ance seal refers to a piston and cylinder with sufficiently

small radial clearance that leakage is negligible without

89
liquid lubricants or piston rings. The piston can be centered

in the cylinder by dynamic magnetic forces controlled by position sensors and servo loops, or by hydrodynamic or hydrostatic

fluid forces. Either way, a piston can be centered sur- prisingly well in the cylinder, with radial clearnces of 25 ym

or less, so that there are no solid surfaces in contact, and

theoretically, no wear. With such small clearances it is

necessary to use materials with good dimensional stability,

and also to use hard materials to avoid disaster in case

some rubbing does occur. Possible materials include chromium

oxide surfacing on aluminum, nitrided steel, certain metal

carbides, and ceramics. (We have made a 50-mm diameter

piston and cylinder of fine-grained aluminum oxide ceramic,

with 4 ym radial clearance, which have passed preliminary

tests but have not yet been used as a compressor. For this

purpose we will use the traditional crankshaft and crosshead

to couple to a drive motor.) By coupling the piston and

reciprocating drive motor to a metal, magnetic, or gas

spring (the latter might be called a bounce chamber or some

such term) , a mechanical resonance can be made to match the

desired drive frequency, so that the drive motor need only

supply the forces inherent in the refrigeration process and

various irreversible effects.

With regard to irreversibility, or lack thereof, G. M.

Benson (private communication) has achieved excellent efficiency with small Stirling engines and refrigerators, partly

through the use of isothermalizers, that is, systems of con- centric interleaving cylindrical rings in the expansion and

compression spaces which provide large heat transfer areas

so that so-called isothermal processes are actually very

nearly isothermal (23)

.

90
Appendix A

Figure 3 is a sketch of a Stirling cryocooler with 5-stage

displacer which maintains temperatures below 7 K. Figure 4 summarizes the operating characteristics of the cryocooler.

The temperature of the cold end is shown as a function of

electrical heat input to an aluminum cylinder 13 by 130 mm

which simulated the SQUID gradiometer later mounted in the

cold space. Temperatures of the first four stages run at

about 167 K, 70 K, 27 K, and 15.5 K. The compressor cylinder

and piston, with 90 cm displacement, were larger than, but

otherwise similar, to those shown previously (12). The crank

and connecting rod mechanism from a small air compressor were

adapted to drive the piston and displacer of the cryocooler

at 1.1 Hz. The relative phase between piston and displacer

motions was adjusted empirically for minimum temperature.

Displacer dimensions in Fig. 3 are given in mm, and other

dimensions, such as wall thicknesses of the glass-epoxy

tubing, are drawn in correct proportions. The displacer is

shown in its lowest position so the compression space is at

maximum volume and the five expansion spaces are at zero volume. The assembly is shown as it should appear at low

temperature. To achieve lowest temperature, it is essential

to allow for differential contraction of the nylon and the

glass-epoxy, to ensure that the lower expansion spaces close

up completely at the bottom of the displacer stroke. It

is less important for the upper expansion spaces to close

completely. We determined that the differential contraction

between room temperature and cyogenic temperatures was almost exactly 1%. We accordingly made the four lower

displacer sections about 1.1% longer, at room temperature,

than the corresponding cylinder sections, so that the lowest

expansion space would go to zero volume at the bottom of the

displacer stroke while the upper expansion spaces would have

successively larger, but still small, finite mimimum volumes.

91
Displace.-

Push Rod Solid Nylon

Stopped Dlaplacer

(In lowest position)

' r Aluminum Spacer

(where required)

Glaae-Epoxy Composite

Cylinder Walla

Thin Aluminum

Radiation Shields

Ahimlnlied Plastic (10-20 layera In vacuum space)

Aluminum End Cap

Cold Space For Device

Vacuum Wall

Fig, Scale drawing of NBS 5-stage cryocooler.

92
90 cm3

7 mm\[Air 9t38x 140

29x130

19x120

9.5x83

4.8x133

5 10 COOLING POWER, mW

h" 0.92 s «-j

UJ

a. /V-5.7x105 Pa S\

Z)

V)V)

UJ

a.a. ~T6x105 Pa TIME

Fig. 4. Displacer and cylinder dimensions (upper left),

temperature vs

.

heat input to cold end (right) , and pressure

variation during two cycles of operation, with short vertical

marks indicating time of displacer bottom-dead-center position

(bottom)

.

With regard to radial dimensions, the four lower displacer

sections were simply cut for an easy push fit (at room

temperature, probably 10 to 20 pm radial clearance.) Differential contraction provides plenty of radial clearance,

(probably more than optimum) when the machine is cold. On

the first (upper-most) section, a uniform taper from about

20 to 100 ym radial clearance, bottom to top, was provided.

It is convenient, for several reasons, to make the displacer

in five individual pieces and pin them together as indicated

in the figure. In particular, it is much easier to fit the

93
individual sections to the corresponding cylinder, and a mistake in one section does not waste the whole displacer.

Also, joints between sections can be made slightly loose to

permit self-alignment of each section in the cylinder.

If a displacer section fits too tightly, that section may

heat up due to friction when the machine is started at room

temperature. The remedy, obviously, is to remove a few urn of material from the offending section. Copper-constantan

thermocouples were used to measure temperatures of the

upper, warmer stages and 1/8-W carbon resistors for the

lower stages.

Some performance data are given in Fig. 4. The vertical

coordinate in the graph of cooling power is the temperature

of an aluminum cylinder in the cold space, serving as a dummy load. During the last week before this conference, a SQUID gradiometer has been operating in the cryocooler.

Performance of the system will have to be left for a later

publication. This work is being done in collaboration with

D. B. Sullivan.

References

1. Letter from the Editor, "Opportunities for Miniature

Cryocooler Research and Development" . Cryogenics 30 , 164 (.1980) . 2. Zimmerman J. E., and Flynn, T., eds . : Applications of

Closed-Cycle Cryocoolers to Small Superconducting

Devices. NBS Special Publication SP-508, Superintendent

of Documents, U. S. Government Printing Office, Washington,

D. C. 20434 (1978) . (Also available from the authors) . 3. Proc. Cryogenic Cooler Conf., USAF Academy, Co., Oct.

1973, Tech. Report AFFDL-TR-73-149 , Vol. 1, Wright- Patterson Air Force Base, Ohio 45433 (Dec. 1973).

4. Radebaugh, R. : See Ref. 2, p. 7.

5. Sherman, A.: Astronautics and Aeronautics 39, Nov. (1978).

94
6. Zimmerman, J. E., and Sullivan, D. B. : Cryoqenics 19,

170, (1979). —

7. Steyert, W. : See Ref. 4, p. 7.

8. Barclay, J. A.: Cryogenics (to be published).

9. Pratt, Jr., W. P., et al. : Cryogenics 17, 689 (1977).

See also Reports LA-8111 and LA-8134, Los Alamos

Scientific Laboratory, Los Alamos, NM 87545.

10. Radebaugh, R. : See Ref. 2, p. 93.

11. Rule, T. T., and Qvale, E. B.: Adv. in Cryogenic

Engineering, K. Timmerhaus, ed. , V14 , 343 (1969),

Plenum Press.

12. Zimmerman, J. E. , and Radebaugh, R. : See Ref. 2, p. 59.

13. Feller, D. W. : Electronic Design 23_, 116 (1974).

14. Zimmerman, J. E. and Campbell, W. : Geophysics, 40, 269

(1975). —

15. Campbell, W, and Zimmerman, J. E.: IEEE Trans. Geosci.

Elect. (1980) (Accepted for publication).

16. Zimmerman, J. E. , et al. : IC SQUID, edited by Hahlbolm

and Lubbig.

17. Sullivan, D. B. , and Zimmerman, J. E. : Int. J. of Refrig.

2, 211 (1979).

18. du Pre, F. K., and Daniels, A.: Prog, in Refrig. Sci. and

Tech., Vol. 1, AVI Publishing Co., 137 (1973).

19. Little, W.: See Ref. 2, p. 75.

20. Little, W. : AIP Conf. Proc. No. 44, edited by Deaver

et al. , (1978)

21. Gasser, et al. , Applications of Cryogenic Technology , Vol, 4, 416 (1971).

22. Benson, G. : U. S. Patent Nos. 3, 928, 974 and 4,044,558.

23. Benson, G.: Proc. 12th IECED, 416 and 1478 (1977).

24. Brenckenridge, R. W. et al. : Technical Report AFFDL-TR-78

-166, Wright-Patterson Air Force Base, OH 45433 (Dec.

1978).

25. Sherman, A., et al.: in Advances in Cryogenic Engineering,

ed. by Timmerhaus, Clark, and Reed, V25 . Plenum Press (to

be published)

95
"Refrigeration for Cryogenic Sensors and Electronic Systems," NBS Special Publication SP-607, (J. E. Zimmerman, U. B. Sullivan, and S. E. McCarthy, eds.), Superintendent of Documents, U. S. Government

Printing Office, Washington, D. C. 20234 (1981), pps. 186-194.

OPERATION OF A PRACTICAL SQUID GRADIOMETER

IN A LOW-POWER STIRLING CRY0C00LERt

D. B. Sullivan, J. E. Zimmerman, and J. T. Ives National Bureau of Standards

Electromagnetic Technology Division

Boulder, CO 80303

A commercial Nb point-contact SQUID has been used

in the construction of a simple magnetic gradiometer

system supported by a split Stirling refrigerator. The

magnetic interference generated by the cryocooler was sufficiently small to permit the acquisition (for demonstration purposes) of a reasonable magnetocardiogram.

Preliminary investigation of the source of the residual

refrigerator interference indicates that cyclic temperature variation is the largest single factor influencing

the gradiometer output. A simple electronic temperature

controller has been added to a different cryocooler to study temperature stabilization. Better than 1 mK

stability has been achieved at 8.5 K.

Key words: Cryocoolers; low temperature; magnetic

gradiometer, refrigeration; SQUID; Stirling cycle;

superconductivity.

1. Introduction

The economic viability of superconducting instrumentation would be greatly enhanced if

properly matched cryocoolers were readily available to support their operation. By properly

matched we mean: 1. cost comparable to or less than the supported instrument, 2. cooling

capacity not significantly greater than that required (often only a few milliwatts),

3. package size comparable to or not much larger than that of the device in a liquid-helium

cryostat, and 4. vibration and magnetic interference levels which do not appreciably affect

the instrument operation. This last requirement is especially critical for SQUID magnetometers and magnetic gradiometers which are designed to sense environmental magnetic fields

or gradients (e.g., geophysical -.fiel ds or biological fields). The SQUID magnetometer has a sensitivity of the order of 10" times the earth's magnetic field and thus extremely small

angular vibrations translate directly into interference if the device is operated in such a field. In this paper we discuss this and other potential interference sources and describe

a SQUID gradiometer supported by a low-power cryocooler [1,2]. As a demonstration of the

level of interference for this system, a magnetocardiogram was recorded.

2. Cryocooler

The cryocooler used in this work has been recently described by Zimmerman [2]. Figure

1, taken from this reference, shows the details of construction of the cold end of this

split Stirling refrigerator. The refrigerator operating at 1 Hz has achieved a temperature

of 7K with no heat load (an aluminum rod occupied the 'cold space for device' shown in

Figure 1). The compressor displacement is 90 cm^ and optimum operating helium pressure is just over 2 atmospheres (gauge pressure). The displacer stroke is 0.7 cm.

+ Contribution of the U. S. Government, not subject to copyright. Work supported by the

Navy Coastal Systems Laboratory.

tt Present address: Engineering Department, University of Utah, Salt Lake City, Utah.

96
isplacer

ush Rod p

Aluminum

Flange

9.5

Vaccum

Wall

Pinned

Displacer

Joint

4.8

z Annular

Expansion

Space

'>•

SF

4- -•

I i

Solid Nylon

Stepped Displacer

(in lowest position)

^Aluminum Spacer

(where required)

Glass-Epoxy Composite

Cylinder Walls

Thin Aluminum

Radiation Shields

Aluminized Plastic

(10-20 layers in vacuum space)

Aluminum End Cap

Cold Space For Device

\\\\\V\VV^ \

Vacuum Wall

Fig. 1. Details of 5-stagev split Stirling cryocooler.

Figure taken from reference [2].

Dimensions are in millimeters.

97
The selection of many of the design features for the cryocooler was influenced by

interference considerations. These design features are described in section 4 below

under the appropriate potential interference sources.

3. SQUID Gradiometer

The gradiometer and its principal dimensions are shown schematically in Figure 2. The

diameter of the gradiometer was selected such that none of the shields of its support re- frigerator (Figure 1) would need modification. This was done to keep the system as simple

as possible, but the resulting gradiometer was not particularly sensitive with the limited

diameter and baseline. The commercial toroidal SQUID is centered within a niobium tube between the end coils. This tube is moved vertically to achieve axial balance. Small superconducting tabs were attached to the tube (in pairs, one at each end) to achieve transverse

balance.

1.59 cm

ALUMINUM PLUG (TAPPED)

FOR CONNECTION TO COOLER

COIL (10 TURNS)

CUTOUTS FOR

ACCESS TO SQUID

EPOXY-GLASSTUBE

COIL (10 TURNS)

Fig. 2. SQUID gradiometer. The SQUID is inside a niobium cylinder.

98
The system was balanced to approximately 1 x 10 both axial ly and in the transverse

direction while operated in liquid helium. The sensitivity was found to be ^ 10 T/m per

flux quantum with a noise level of ^ 10" T/m in a 40 Hz bandwidth. For thermal anchoring,

four strips of aluminum foil (4 mm wide by 0.3 mm thick) were tied along the entire length

of the gradiometer with the ends also tied to the aluminum cold end of the refrigerator.

The SQUID was also independently anchored (thermally) to the cold end. The gradiometer

coils (10 turns each) were connected in parallel to the SQUID rather than the usual series

connection. This was done so that more turns could be used on each coil and still have the

combined inductance match that of the SQUID coil.

For the interference tests the gradiometer was operated partly within an aluminum

shielded room [3], as shown in Figure 3. The shielded room provides some isolation between

compressor and displacer since its cutoff frequency is 1/3 Hz. The cylinder with displacer

was mounted rigidly to the roof of the room to avoid any rocking motion induced by the

rotation of the displacer drive shaft (see Fig. 3). This rigid mounting led to additional

microphonic noise in the 3 to 10 Hz range. Vibrations in this frequency range were generated

by some heavy refrigeration equipment in the building.

COMPRESSOR AND MOTOR

DISPLACER DRIVE SHAFT

•HELIUM GAS LINE CRANKSHAFT

'^yy

DISPLACER AND CYLINDER

IN VACUUM JACKET

ALUMINUM ROOM

Fig. 3. Mounting of the cryocooler in an aluminum shielded room.

99
4. Potential Sources of Interference

In previous papers we emphasized the refrigeration aspects of the cryocooler while alluding to certain characteristics which would minimize the magnetic signature of the device.

In this section we expand upon the design features relating to magnetic interference.

4.1. Motional Interference

The use of low speed (in the neighborhood of one Hz) and the low power requirement re- sult in low vibrational noise, because of the low static and dynamic forces involved. With

the plastic regenerator materials, this low-speed is, in fact, essential since much higher

speed operation reduces the thermal penetration depth and thus results in regenerator inefficiency, particularly in the bottom stage.

The displacer requires very little power for operation, so the mechanical components

which produce the displacer motion can be very smooth in operation. The compressor and

motor are coupled to the cold portion of the cooler through a drive shaft and a tube through

which the helium gas is transported between compressor and displacer sections. It is es- sential that vibration transmission through these two connections be minimized. To this

end we use ratner light weight and flexible materials (thin-wall brass for the drive shaft

and small diameter copper tube for the helium gas) for these components. By separating the

compressor and displacer one meter from one another we find little vibration transmission

from compressor to displacer. Additional flexibility and versatility might be achieved

through use of a simple hydraulic system to transmit the mechanical motion from the motorcompressor to the displacer.

The one component of motion which still remains as a potential problem is that of

pressure flexing of the displacer cylinder. It is unlikely that this cylinder is perfectly

uniform around its circumference and the cyclic pressure variations could therefore result

in some cyclic transverse motion of the cold end. This problem would be reduced by operation

at the lowest possible pressure.

An earlier paper [1] described the results of operation of a simple SQUID magnetometer in a four-stage cooler. The peak-to-peak signature at the cooler frequency of lrHz was 0.5 nT (SQUID operated in the earth's field). This corresponds to a motion of 10

" radian if this is the only interference source, emphasizing the extreme stability required

by this type of instrument. This field oscillation was dependent upon the orientation of

the SQUID in the horizontal plane, so we concluded that at least part of that signature was due to flexing of the displacer cylinder presumably induced by the pressure oscillations.

There was no direct evidence for other interference mechanisms, but they could easily have

been present as well

.

For magnetic gradiometers, motional interference is often not as severe. If the earth's

field is uniform in the vicinity of the device (the usual case), the motional signature is reduced by the degree of balance of the gradiometer. That is, it is the remnant magnetometer response which picks up the field motion. In this particular case the gradiometer is balanced to a part in 10 so motional inteference is reduced by the same factor.

4.2 Compressor Interference

It is necessary to minimize Hirect magnetic interference by the compressor at the cold

end. Two precautions are taken. First, the compressor Is designed with a minimum of

ferromagnetic components. Materials utilized included brass, aluminum, stainless steel,

and nylon. We have used bronze and nylon bushings rather than ball and roller bearings

which are commonly ferromagnetic. The piston is nylon in a stainless tube. The motor and

gearbox selected for the machine has cast aluminum housing components and appears to have

only a modest magnetic signature (this motor was once operated within 15 cm of the gradiometer with no severe interference). The second precaution involves modest separation (1/2

to 1 meter) of the motor-compressor and the displacer sections of the refrigerator. It is conceivable that some sort of shielding could be added to the motor and compressor to further reduce their magnetic signature, but this does not yet appear to be a major interference source.

100
4.3. Magnetism of Cold End Materials

Plastics were chosen for the critical cold parts of the refrigerator because they are nominally nonmagnetic. However, at the sensitivity level of the SQUID, the magnetic properties of virtually all materials are not negligible. Small remanent moments and magnetic

susceptibilities often result in readily measured effects. The materials concerns fall

into two categories, magnetic interference from moving components and slow drifts which

arise from temperature dependent magnetic effects in materials which are fixed relative to the SQUID.

In the first instance two phenomena are important. The presence of a remanent magnetic

moment in the displacer leads to an obvious magnetic signature. In general, remanent

magnetism in such materials is associated with traces of ferromagnetic substances either

introduced into the bulk of the material during manufacture, or quite commonly, by machine

work during shaping of the component. The latter contamination can be reduced sharply by

careful cleaning procedures. Remaining permanent moments can be reduced by demagnetizing

procedures. Even without remanent magnetism, moving components can generate a magnetic

signature since the ambient field (often the earth's field) induces a moment in anything

with a non-zero magnetic susceptibility.

It is difficult to be quantitative in the case of remanence, but we can estimate the

size of the induced moment at the bottom end of the nylon displacer. Assuming for the

moment that the lowest part of the displacer produces the strongest effect, we suggest that

the displaced volume in the bottom stage is the relevant volume of nylon to consider. That

is, we assume that the effect of cyclic axial motion of a semi-infinite rod in a magnetic

field is the same as the cyclic appearance of a volume of the rod equal to the cross sectional

area times the stroke. Thus for the 5-stage cooler of figure 3, we have a volume of 0.7 x

tt x (2.4) 2 cnA For nylon with a susceptibility of 10"^ in a field of 5 x 10~4 T this yields

a cyclic induced moment of ^ 5 x 10"'^ Am?. This translates to an interference gradient

field of approximately 10"^ T/m. (The spacing between the end of the displacer and the top

coil of the gradiometer is ^ 3 cm). This is well below any interference level which might

be achieved in the near future.

The magnetic susceptibility of many materials can be a rather strong function of

temperature leading to long term drift in the induced magnetic moment. The refrigerator

components most li.kely to experience such temperature drift are the radiation shields and

mylar superinsulation. Drifts of many degrees C in temperature are likely since radiation

heat transfer is very small and the time constant for temperature equilibrium might then be

extremely long. Lacking specific knowledge of either the magnitude of sucti temperature

variation or the temperature dependence of the relevant material susceptibilities, it is not practical to dwell on this question. Needless to say, the appearance of drift might

eventually lead us to attemp to clarify this situation.

4.4. Temperature Variation

Besides this last effect (induced magnetism which drifts with temperature), there is a more direct influence which temperature change can have on a SQUID. If the critical current of the SQUID is a function of temperature, then temperature change will have an effect

on the operating point of the SQUID. In principle this should not change the output of a SQUID operated in a phase-locked loop, but measurements indicate that this is not the case. This relation between SQUID output and operating point (rf level or critical current) is probably a second order effect related to some asymmetry in the interference pattern. It's presence means that temperature changes (either drift or oscillatory changes) can affect

the SQUID output directly. The cures for this problem are twofold. First eliminate the

temperature variations and second, eliminate (or reduce) the temperature dependence of the

SQUID's critical current. Probably both solutions should be considered.

5. Interference Measurement

The gradiometer sensed an interference signal at 1 Hz (compressor frequency) with an amplitude of ^ 6 x 10"^ T/m. We conducted several experiments to isolate the source of

the noise. First, we placed a high permeability shield around the SQUID and found no

101
observable change in the amplitude of the signature. This immediately eliminates compressorgenerated magnetic signals as the predominant interference source. Furthermore, this shield

attenuated the earth's field by at least a factor of 100. The lack of any change in Interference with this shielding suggests that both gradiometer motional noise and induced moments

(non-zero susceptibility) are also unimportant at this level. Finally, we noted rapid drift

when the cooler was turned off momentarily. During such shutdown the SQUID remained fully

in operation. This suggests a strong dependence of SQUID output on temperature as discussed

above. From direct measurements we find that the temperature of the bottom end varies by at least several mK at the refrigerator frequency and previous measurements on the SQUID Indicated a strong dependence of critical current on temperature. Thus we conclude that the

primary source of the apparent, magnetic signature at this level is temperature fluctuations

at the SQUID. This is also consistent with the observation that the signal amplitude 1s

somewhat dependent on temperature since the slope of the I vs. T curve is a function of T and some dependence is thus expected. c

We also noted that the very low frequency noise (less than 1 Hz) was small. After

achieving temperature equilibrium we noted a rather stable signal (less than 2 x 10-11 T/m

variation over a period of three minutes). This could probably be improved through temperature regulation which will be discussed in the next section.

6. Interference and Drift Reduction

We usea a very simple expedient to cancel most (90%) of the 1 Hz signature. This was

to attach a small magnet to the shaft of the compressor so as to generate a field which just

canceled the observed signature. The proper way to deal with this sort of interference re- duction is to use a commutative filter with a reference derived from the compressor shaft.

The advantage of such a filter is that it can also handle harmonics of the fundamental

interference which are clearly present. Nonetheless, the simple cancellation of most of

the interference allowed the following measurement.

As a demonstration of the system we recorded two magnetocardiograms (see Figure 4). The first was taken with the gradiometer in a simple cryostat and the second with the

gradiometer supported by the 5- stage cooler. Bandwidth for the measurements was 2 to 40

Hz (to further reduce the one Hz signature). The added noise in the 3 to 10 Hz range 1s

probably a result of the rigid mounting to the shielded room (see Section 3).

b)

IN CRYOSTAT

a

IN CRYOCOOLER

Fig. 4. Magnetocardiograms taken with gradiometer in a simple cryostat (a) and In the

5-stage cryocooler (b).

102
This result provides a fairly clear indication that mechanical refrigerators can be

compatible with SQUID instruments, altnough tiiere is room for much improvment. Obviously,

the problem of temperature fluctuation and drift must be addressed first. (The reduction of

SQUID temperature sensitivity will also help, but we assume that complete elimination of

such temperature sensitivity is very difficult).

We have experimented with two means for stabilizing the temperature of the cryocooler.

These experiments have been carried out with a 4-stage Stirling cryocooler, the bottom end

of which is shown in Figure 5. The carbon-resistance thermometer and the heater were used

as sensor and control elements in a simple electronic temperature regulator (servo system).

This system was quite effective in eliminating drift. Over a period of two hours, the

temperature change was no more than a fraction of a mK with the temperature at 8.5 K. Because of the rather slow thermal response, the simple regulation circuit did not significantly reduce the 2 mK oscillation (at 1 Hz) which was observed without the regulator.

Further improvement of this regulation is clearly needed. One should note that alternating

current at a frequency above the SQUID sense band can be used for both the heater and thermometer to eliminate any possible interaction between the SQUID and a direct current in these

elements.

NYLON DISPLACER

EPOXY-GLASS CYLINDER

EXPANSION SPACE

ALUMINUM END PLUG

GAS THERMOMETER

BULB (BRASS)

HEATER (10 kQ)

/

CARBON RESISTANCE

THERMOMETER

(220 Q NOMINAL)

BRASS BLOCK

Fig. 5. Details of the bottom end of the 4-stage cryocooler with thermometer and heater

for electronic temperature control. The helium gas bulb permits control of the

thermal mass of this bottom stage.

103
The helium gas bulb shown in Figure 5 provides a simple means for changing the heat

capacity of the cold end of the refrigerator. Even though the volume of this bulb is only

1 cm3, the total heat capacity of the bottom stage could be increased by a factor of 10 by

pressurizing it with helium gas to only 4 or 5 standard atmospheres. This results in a ten- fold reduction in the oscillatory temperature variation and significantly reduces drift even without the electronic temperature regulator.

7. Summary

In summary then, a SQUID gradiometer supported by a split Stirling cryocooler has

been shown to operate at an interference level sufficient to obtain a magnetocardiogram of

modest clarity. Temperature fluctuations produce the primary interference signature and

methods for temperature stabilization have been studied. We expect that with each improvement which reduces magnetic interference, we will identify other interference sources which

affect the system at a lower level. The most logical approach is to proceed to deal with

these as they are uncovered.

8. References

[1] J. E. Zimmerman and R. Radebaugh in "Applications of Closed-Cycle Cryocoolers to Small

Superconducting Devices," NBS Special Publication SP-508, J. E. Zimmerman and T. M. Flynn, eds.), Superintendent of Documents, U. S. Government Printing Office,

Washington, D. C. 20234 (1978). (Also available from the authors), p. 59.

[2] J. E. Zimmerman in "IC-SQUID (Conference)" Berlin, West Germany, May 1980 (to be

published).

[3] J. E. Zimmerman, J. Appl . Phys. 48, 702 (1977).

10A
"Refrigeration for Cryogenic Sensors and Electronic Systems," NBS Special Publication SP-607, (J. E. Zimmerman, U. B. Sullivan, and S. E. McCarthy, eds.), Superintendent of Documents, U. S. Government

Printing Office, Washington, D. C. 20234 (1981), pps . 173-177.

MEASUREMENT OF THERMAL PROPERTIES OF CRY0C00LER MATERIALS

+

J. E. Zimmerman, D. B. Sullivan, R. L. Kautz, and R. D. Hobbs

National Bureau of Standards

Electromagnetic Technology Division

Boulder, CO 80303

Materials used in several low-power cryocoolers are characterized by very low thermal conductivities and moderate to large heat capacities at low temperature. Consequently, thermal

equilibrium times are inordinately long for measurements of

thermal properties to be made by the usual quasi -steady-state

methods. We have developed a method of measurement in which

the thermal conductivity and the specific heat are derived from

observations of the time-dependent temperatures at two or more

points on a cylindrical sample in response to a step function

of heat applied to one end of the sample. Some preliminary

results have been obtained on G-10 epoxy laminated tubing.

These transient measurements give results in much less time

than the steady-state method, and have the. additional advantage

of being adaptable, with fair accuracy, to the cryocooler

structure itself, or to a separate structure geometrically

similar to the cryocooler.

Key words: Composites; cryocoolers; plastics; refrigeration;

regenerators; specific heat; thermal conductivity.

I. Introduction

Measurement of thermal properties of plastic and composite materials used in the con- struction of several low-power-cryocoolers requires long times if done by conventional

methods. For example, a sample of nylon whose longest dimension is 5 cm would require

something like two hours to reach thermal equilibrium after an input of heat at one end.

This estimate is based on a thermal conductivity K^ 0.5 mW/cm K and a specific heat c ^

25 mJ/cm K at 10 K. Thermal equilibrium times will of course be proportionately greater

for special composite materials with large heat capacities. We have therefore developed

a method of measurement in which we apply a step function in heat flow to one end of a cylindrical sample and observe the propagation of the transient in temperature as the heat

diffuses along the sample. Thus, thermal properties of the sample material are derived

from measurements made during the initial development of the transient, rather than after

the transient has died out as in conventional equilibrium or steady-state ac methods[l].

This gives an order of magnitude reduction in time required for the measurements, not

to mention the fact that both the thermal conductivity and the specific heat are derived

from the same set of measurements. In addition, the method can be used for rough measurements on the cryocooler structure itself, the geometry being appropriate.

tContribution of the U. S. Government, not subject to copyright. Work supported by the

Office of Naval Research.

ttPresent address: Physics Department, Indiana State University, Terre Haute, Indiana.

105
Transient measurements of thermal properties have been made previously on cylindrical

samples by a heat pulse method. Our reasons for using a step function rather than a pulse was that an Ideal heat pulse momentarily raises the end of the sample to a high temperature,

thereby enhancing the possibility of heat loss by extraneous mechanisms such as radiation,

and also complicating the analysis through the temperature dependence of the thermal con- ductivity and specific heat. The magnitude of the step function can be adjusted to give

a slow, steady, and relatively small, increase in temperature of the end of the sample.

II. Analysis of the Method

The Ideal geometry 1s a uniform semi-infinite bar with a heater on the free end, x = 0,

and thermometers to measure temperatures T, and T~ at two points x, and x~ (Fig. 1). The

simplest initial condition is to let the bar come to thermal equilibrium at a temperature T , and then at time t = 0, turn on the heater at a power level P. The one-dimens1onal diffusion

equation for this situation is

* 2T

ax

2 Hat

(1)

where k and C are respectively the thermal conductivity and specific heat per unit volume

of the material. The solution for the stated initial condition Is:

T(t,x) = T +

2P

A/iFkC

/Fe -Cx^/(4kt) . P_ xerfc [<^>

1/2

]

•

(2)

where erfc means the complementary error function, and A 1s the cross-sectional area, solution is plotted as a function of x and of t in Fig. 2.

This

/~

HEATER *s

THERMOMETERS^

3

Fig. 1. Experimental arrangement for thermal properties measurement. The heater 1s turned on abruptly at t = (step function) and the resultant temperature

variation (sensed by the two thermometers) provides the means for determining

the thermal conductivity and specific heat. The right hand end is connected

to a heat sink, the temperature of which is controlled to set the equilibrium

value of the temperature at T . The sample 1s suspended in a vacuum and

surrounded by a liquid helium cooled shroud to minimize heat leaks to it. The

leads from the carbon resistance thermometers are heat sunk to this shroud for

similar reasons.

106
TM"To

(a)

^5 (TIWl-TJ

DISTANCE. K/uJkUc)

20

101 (b)

o

1 2 3 .4 .L. — ?-

100 200 300 400 800 000

TIME. t(4k/Cx2

)

Fig. 2. Spatial (a) and temporal (b) solutions of diffusion equation.

7700

107
Our method of measurement is to record the temperatures, T, and T^ as a function of time. This gives a pair of curves of the shape of Fig. 2b, but 'with different time scales since Ti changes more rapidly after the power is turned on than L>, and of course L, is always less than T

]

. From these curves we can read off a pair of numbers T, and L, for a particular time t'. For that particular time we then have both coordinates (T, , x?) and

(T^.x^) for two points on the curve of Fig. 2a, which establishes the horizontal and

vertical scales of that curve. This process involves use of the ratios x,/x

? and T,/T? and

an interactive procedure (using a computer) which locates the two unique points having the

correct values of these ratios. The ratio, k/C, is obtained directly from the horizontal

scale factor and the product kC can be derived at x = from the vertical scale factor and

equation (2). The combination of this ratio and product yield the values of k and C.

It is only essential, of course, to measure T, and T

? at some convenient time t. How- ever, measuring T, and Tp as a function of time provides a high degree of redundancy which

is useful in checking for errors and inconsistencies in the method.

Since k and C are usually temperature dependent at low temperatures, it is important to choose a power level and time scale such that T, and T~ do not rise ^ery much above the

initial value T . Otherwise, the temperature dependence of k and C must be considered

in equation (2), making the analysis difficult if not impossible. The use of small temperature intervals is standard practice in specific heat and thermal conductivity measurements.

Since the bar is not semi-infinite, the measurement time t' should be chosen small

enough that no appreciable amount of heat reaches the far end of the bar, otherwise the

spatial and temporal profiles given in Fig. 2 will no longer be accurate. Since the

spatial profile, Fig. 2a, approaches zero very quickly at large x, it seems satisfactory

to choose xo ( tne position at which Tp is measured) at about the middle of the real bar,

and then choose t

1 such that the rise in Tp is no more than about 25% of the rise in T,

.

It is assumed that the position x, should Be near the free end of the bar. Any deviation

of the temporal profile from the curve of Fig. 2b might indicate that one or more of these

conditions has been violated.

6.0 r5.5 -

5.0 -

4.5 20 40 60 80 100 120

TIME (SECONDS)

Fig. 3. Temperature as a function of time for two thermometers on a G-10 epoxy fiberglass

rod. The thermometers were 1.2 cm and 4 cm from the heater. The rod had a cross- sectional area of 0.895 cm . 108
III. Results

Figure 3 shows the temperature at 2 thermometers on a G-10 epoxy-fiberglass rod as a . function of time after a step function was applied to the heater. The rod area was .895 cm and the thermometers were located 1.2 cm and 4.0 cm from the heater.

We found that the consistency of the results was poor if we selected a time t for

which the temperature change at the second thermometer was small. This is presumably

related to problems of resolution in the temperature at this point. We also found that

the results varied considerably if the heat wave was allowed to reach the heat sink (the

temperature controlled end of the bar) before the temperature points were taken. This

clearly relates to a breakdown in the assumption that the rod was semi -infinite.

For the data of Fig. 3, reasonable consistency in results was achieved over the

time interval of 40 to 75 seconds. While the temperature differences are a bit larger

than desired, we used the method described in the previous section to determine the

thermal conductivity and specific heat of this material at 5 K., The thermal conductivity

was found to be 1.1 mW/cm-K and the specific heat was 4.7 mJ/cm -K.

These numbers fall within the range of values given by other investigators. Additional

work must be done to arrive at a quantitative criterion for selection of valid data reqions.

The measurement scheme is to be automated so as to provide for rapid acquisition of information over a modest temperature range for a variety of materials. One should note that this

technique has been devised to deal with rapid evaluation of a wide range of materials. An accuracy of 10-20% should suffice for such a study and this method seems well suited to the

task.

IV. References

[1J. G. C. Danielson and P. H. Sidles, Thermal Diffusivity and Other Non-Steady-State

Methods, in Thermal Conductivity, edited by R. P. Type, 149-200 (Academic Press,

New York, NY, 1969).

109
NBS-114A (REV. 2-8C) U.S. DEPT. OF COMM.

BIBLIOGRAPHIC DATA

SHEET (See instructions)

1. PUBLICATION OR

REPORT NO.

NBS TN-1049

2. Performing Organ. Report No 3. Publication Date January 1982

4. TITLE AND SUBTITLE

A Study of Design Principles for Refrigerators for Low-Power Cryoelectronic Devices

5. AUTHOR(S)

J. E. Zimmerman and D. B. Sullivan

6. PERFORMING ORGANIZATION (If joint or other than NBS, see instructions;

NATIONAL BUREAU OF STANDARDS

DEPARTMENT OF COMMERCE

WASHINGTON, D.C. 20234

7. Contract/Grant No.

8. Type of Report & Period Covered

9. SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (Street, City, Stote, ZIP)

10. SUPPLEMENTARY NOTES

"2 Document describes a computer program; SF-185, FlPS Software Summary, is attached.

11. ABSTRACT (A 200-word or less factual summary of most significant information. If document includes a significant

bibliography or literature survey, mention it here)

This report summarizes a five-year effort at NBS which has been directed

toward the development of low-power cryocoolers suited to the support of superconducting instruments. The report deals with a variety of aspects of construction

and operation of refrigerators as well as with a model which allows one to optimize

the design for minimum drive power. The publications generated by the program

are included as an appendix.

12. KEY WORDS (Six to twelve entries; alphabetical order; capitalize only proper names; and separate key words by semicolons)

Cryocooler; cryogenics; low temperature; refrigerator; Stirling cycle;

superconducting devices.

13. AVAILABILITY

fXf Unlimited

|

| For Official Distribution. Do Not Release to NTIS

~\ Order From Superintendent of Documents, U.S. Government Printing Office, Washington, D.C.

20402.

[~Xl Order From National Technical Information Service (NTIS), Springfield, VA. 22161

14. NO. OF

PRINTED PAGES

114

15. Price

$11.00

•frU.S. GOVERNMENT PRINTING OFFICE:1982—577-445 ( 113 USCOMM-DC 8043-PSO
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