A111D3 DTBETE

NATL INST OF STANDARDS & TECH R.I.C.

A11 103073272

/Refrigeration for cryogenic sensors and QC100 .U57 NO.607, 1981 C.I NBS-PUB-C 19 NBS SPECIAL PUBLICATION 607

U.S. DEPARTMENT OF COMMERCE / National Bureau of Standards

Refrigeration for Cryogenic Sensors

and Electronic Systems

r Sponsored by

International Institute of Refrigeration—Commission A 1 /2

rnj Office of Naval Research—Naval Research Laboratory

DU/ Cryogenic Engineering Conference

National Bureau of Standards
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or iXAKVAsm

Refrigeration for Cryogenic Sensors

and Electronic Systems

JUN 1 5 1981

/la--

no- ^o^f

Proceedings of a Conference held at the

National Bureau of Standards,

Boulder, CO, October 6-7, 1980

Edited by:

J. E. Zimmerman, D. B. Sullivan, and S. E. McCarthy

Electromagnetic Technology Division

Center for Electronics and Electrical Engineering

National Engineering Laboratory

National Bureau of Standards

Boulder, CO 80303

Sponsored by: International Institute of Refrigeration — Commission A 1/2

Office of Naval Research — Naval Research Laboratory

Cryogenic Engineering Conference

National Bureau of Standards

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

NATIONAL BUREAU OF STANDARDS, Ernest Ambler, Director

Issued May 1981
Library of Congress Catalog Card Number: 81-600038

National Bureau of Standards Special Publication 607

Nat. Bur. Stand. (U.S.), Spec. Publ. 607, 223 pages (May 1981) CODEN: XNBSAV

U.S. GOVERNMENT PRINTING OFFICE

WASHINGTON: 1981

For sale by the Superintendent of Documents, U.S. Government Printing Office, Washington, DC 20402

Price $6.50

(Add 25 percent for other than U.S. mailing)
ABSTRACT

This document contains the proceedings of a meeting of refrigeration specialists held at the National Bureau of Standards, Boulder,

CO, on October 6 and 7, 1980. Participation included representatives of industry, government, and academia. The purpose of the

meeting was to discuss progress in the development of refrigeration

systems which have been specialized for use with cryogenic sensors

and electronic systems. The meeting focused primarily on the

temperature range below 20 K and cooling capacity below 10 W. The meeting was jointly sponsored by the International Institute

of Refrigeration-Commission A 1/2, the Office of Naval Research,

the Naval Research Laboratory, the Cryogenic Engineering Conference,

and the National Bureau of Standards.

Keywords: Cryocoolers; cryogenic sensors; helium; refrigeration;

superconducting devices.

DISCLAIMER

Except where attributed to NBS authors, the content of individual sections of this

volume has not been reviewed or edited by the National Bureau of Standards. NBS therefore accepts no responsibility for quality of copy, comments, or recommendations therein.

The mention of trade names in this volume is in no sense an endorsement or recommendation

of the National Bureau of Standards.

1 1

1

TABLE OF CONTENTS

Page

Cover: Drawing of the first Stirling engine from the original 1816 patent. The process of

regenerative heat exchange in this engine is a key feature of many present day cryocoolers. From the book Stirling-Cycle Machines by G. Walker (Clarendon Press, Oxford,

1973).

1. Introductory Remarks and Summary, 1

2. "A Preview of Unconventional Suggested Cooling Processes and Some

Practical Problems Related to Small Refrigerators,"

by Pierre M. Roubeau 3

3. "Refrigeration Requirements for Superconducting Computers,"

by R. W. Guernsey and E. B. Flint 15

4. "Development Approaches for Long-Life Cryo- Coolers,"

by Ronald White and William Haskin 21

5. "Theoretical Analysis of A 3-Stage Stirling Cycle Cryocooler,"

by Stuart B. Horn and Mark S. Asher 30

6. "Some Thermodynamic Considerations of Helium Temperature Cryocoolers ,"

by D. E. Daney 48

7. "Dynamic Analysis of a Small Free-Piston Resonant Cryorefrigerator,"

by R. A. Ackermann 57

8. "Methods for the Measurement of Regenerator Ineffectiveness,"

by Ray Radebaugh, Del Linenberger, and R. 0. Voth 70

9. "Serviceable Refrigerator System for Small Superconducting Devices,"

by Ralph C. Longsworth 82

10. "Performance of A 1 Watt 4K Cryosystem Suitable for A Superconducting Computer,"

by E. B. Flint, L. C. Jenkins, and R. W. Guernsey 93

11. "Developments Toward Achievement of a 3-5 Year Lifetime Stirling Cycle

Refrigerator for Space Applications,"

by Max G. Gasser, Allan Sherman, and William Beale 103

12. "Thor Cryocooler,"

by G. M. Benson and R. J. Vincent 116

V
TABLE OF CONTENTS, continued

13. "A Cryogenic System for the Small Infrared Telescope for Spacelab 2,"

by E. W. Urban, L. Katz, J. B. Hendricks, and G. R. Karr 117

14. "Requirements for and Status of a 4.2°K Adsorption Refrigerator Using Zeolites,"

by William H. Hartwig 127

15. "Cooling of SQUID Devices by Means of Liquid Transfer Techniques for Reduced

Helium Consumption and Enhanced Temperature Stability,"

by El don A. Byrd and R. G. Hansen 136

16. "A Contamination Free Compressor for Small Scale Stirling Refrigerators,"

by J. G. Daunt and C. Heiden 141

17. "A Bi-Directional Linear Motor/Generator with Integral Magnetic Bearings

for Long Lifetime Stirling Cycle Refrigerators,"

by Philip A. Studer and Max G. Gasser 146

18. "Destgn Considerations for Mixromintature Refrigerators Using Laminar Flow

Heat Exchangers,"

by W. A. Little 154

19. "Progress in the Development of Microminiature Refrigerators Using

Photolithographic Fabrication Techniques,"

by R. Hollman and W. A. Little 160

20. "Low Temperature Regenerators,"

by Lawrence A. Wade

21. "Measurement of Thermal Properties of Cryocooler Materials,"

by J. E. Zimmerman, D. B. Sullivan, R. L. Kautz, and R. D. Hobbs 173

22. "Gas Heat Switches,"

by Emanuel Tward . i/o

23. "Operation of a Practical SQUID Gradiometer in a Low-Power Stirling Cryocooler,"

by D. B. Sullivan, J. E. Zimmerman, and J. T. Ives

24. "Current Status of High Temperature Josephson Device Technology,"

by M. Nisenoff

210

25. Attendance List

vi
INTRODUCTORY REMARKS AND SUMMARY

This document contains the proceedings of the second of two conferences held at the

National Bureau of Standards Boulder Laboratories on the subject of closed-cycle cryocoolers for small superconducting devices and electronic systems.

The first conference, whose proceedings are available from the Superintendent of

Documents, U. S. Government Printing Office, Washington, DC, as NBS Special Publication

508, entitled "Applications of Closed-Cycle Cryocoolers to Small Superconducting Devices"^

was a meeting of about 40 invited speakers and participants. Its purpose was to review

the state-of-the-art of small cryocoolers for temperatures below. about 20 K, and to describe the needs of various prospective users of cryocoolers for superconducting and

other cryogenic instruments used in biomagnetism, geophysics, magnetic anomaly detection,

superconducting computers, Josephson voltage standards, space applications, millimeter-wave

and infrared detection, and laboratory measurements. Also presented at this meeting were

some new ideas for very- low-power low- interference cryocoolers, Stirling and Joule-Thomson

types, particularly suited to many of the listed applications. A subject of special

interest was a review of progress on high-T superconductors, since the cost, complexity,

and drive power of a cryocooler can be reduced considerably if the instrument can be

operated at temperatures above 6 or 8 K rather than at 4 K or below. There are, of course,

some applications where lower temperatures are required in any case.

In short, the purpose of the first conference was to define the problem. The present

conference was organized quite differently. It was open to all interested participants,

and most papers were contributed. Invited papers are those by Roubeau and by Nisenoff,

the former being a rather free-ranging discussion of some unconventional approaches to cryogenic technology, and the latter a comprehensive review of recent developments in highT superconducting devices and materials. The contributed papers are on refrigeration concepts, systems and components suited to the support of cryogenic sensors and electronic

systems below 20 K. Techniques and components developed for higher temperatures that are

applicable to low-temperature refrigeration are included.

In reviewing the progress that has been made during the last three years, it is

apparent that the desired goals have not been achieved. Nevertheless, there have been some

notable developments. Several funding agenices, in particular the Office of Naval Research

and Wright-Patterson Air Force Base, are now supporting conceptual studies and developments

of low-power cryocoolers for superconducting instruments, and a number of such projects

are underway. As described in this report, a helium 1 iquefier-cryocooler for a superconducting computer has been built and is being tested. Also in this report are papers demonstrating increasing interest and emphasis on gas refrigerators using resonant mechanical

systems driven by linear reciprocating electric motors and supported by gas bearings or magnetic suspensions. These systems offer the potential of ultimate simplicity, freedom

from contamination, long life, and (eventually) low cost. Although the operation of such

systems may at ftrst glance appear somewhat subtle, we might point out the obvious, namely

that the mathematical description of a resonant reciprocating cryocooler is identical (except for non-linearities) in all essentials to that of coupled electrical resonant circuits

with a sinusoidal driving force, a system as familiar as sunrise to most electronic engineers

(Probably even the non-linearities are analogous to a considerable extent). Finally, a book

by Prof. Q. Walker, Stirling Machines, has just been published by Plenum Press, and a companion book, Cryocoolers, is in press. These books are a very comprehensive source of

information and references on the whole technology. •

1
Approximatley one-half (57) of the participants at this meeting returned a questionnaire, distributed during the meeting, with answers to three questions. To the question

"Has the conference been worthwhile?", 56 said yes and 1 said somewhat. To "Should it

be held again?", the answer was yes, unanimously. To "At what interval?" 11 said 1 year,

33 said 2 years, 12 said 3 years, and 1 said 4 years.

There were several useful comments and criticisms. One was to allow more lead time

and to provide for distribution of reprints at the meeting. Some comments concerning more

industry participation are well taken and we suggest that an earlier planning meeting by ! an organizing committee made up of an even mix of industry, university, and government

representatives might lead in the right direction. One should note that both government

and iTidustry papers were withdrawn because they could not be cleared by their organizations

[proprietary considerations will certainly constrain some industry participation). We

beli^eve that the conference will be more lively and useful if the mix of participants

provides for dtscussion of both new concepts and well conceived and engineered systems.

There were several suggestions that the papers should have been better screened. This|

is an oft-raised criticism with every conference. It is clear that papers can be rejected

for obvious technical flaws or for failure to fit into the basic limits of the conference.

However, judgments concerning the quality of a contribution are very difficult to make

and many conferences have opted to exercise such judgment only in the published conference

proceedings. This is what we chose to do in 1980. The next organizing committee will have

to face the question again.

To conclude, and quite apart from the question of the usefulness of this meeting or similar meetings, it might be useful to repeat again the growing conviction that

the ultimate fate of superconducting and other cryogenic instruments will depend heavily

on the development of reliable, compatible, low-cost cryocoolers like those envisioned

by the aurhors of the following papers.

The Editors

Organizing Committee J.

E.

R.W.M.

J.A.

D.

K.

E. Zimmerman, National Bureau of Standards - Chairperson

A. Edelsack, Office of Naval Research

W. Guernsey, International Business Machines

Haskin, Air Force Flight Dynamics Laboratory

Nisenoff, Naval Research Laboratory

L. Olsen, Eidg. Technische Hochschule

Sherman, National Aeronautics & Space Administration

B. Sullivan, National Bureau of Standards

D. Timmerhaus, University of Colorado

Conference Secretary S. E. McCarthy, National Bureau of Standards

I
A PREVIEW OF UNCONVENTIONAL SUGGESTED COOLING PROCESSES AND SOME

PRACTICAL PROBLEMS RELATED TO SMALL REFRIGERATORS

Pierre M. Roubeau

DPh-G/PSRM - CEN Saclay

B.P. N''2 - 91190 Gif-sur-Yvette, France

This paper describes two non-conventional means of refrigeration,

perhaps out of reach of practical realization, whose purpose is to shed

light on some properties of gases and adsorbants having great potential

applications. An account is given of a variation of the Stirling machine

and finally a sketch is made of a crossed cycles method independent of

the refrigeration process used.

This descriptive part is sandwiched by overall considerations on heat

losses, mainly by the cryostat necks and in general by any structures

for which the non-heat conducting properties could, in fact, be an important criterion of success for cryocoolers. Concerning minicoolers which

make use of a circulating coolant, the need for a dry and tight compressor

is equally emphasized and some solutions are suggested.

Key words: Adsorption; cryocoolers; cryostats; helium; refrigeration;

Stirling cycle.

1. Foreword

In the following, the refrigeration process will be considered between essentially room

and liquid helium temperatures. Transposition to another fluid and another temperature range

would in general remain valid subject to proper numerical adjustments. .

2. No I to liquid helium

The straightforward means of cooling a system down to liquid helium temperature is,

evidently... to immerse it in liquid helium.

But . .

.

In the first place this is an expensive liquid. However, the purpose of the "cryocoolers",

topic of this meeting is not only to save money, as it is well known that the scale factor is

dramatic for small installations. At the refrigerator itself the negative calorie will always

be cheaper the bigger the plant.

The problem is

1) most of the users are not nearby big liquefiers,

2) liquid helium is costly to transport,

3) its conservation is expensive and, even for relatively short durations, it is practically

impossible to conserve in small quantities^

4) finally only very competent and conscientious personnel are able to handle it without

prohibitive spillage ; and when, by chance, such people are available it is more advantageous to assign them to more productive tasks.

3
Evidently, other parameters come in balance, such as need of independence, weight, and

volume considerations, available finance etc...

In every case the final conclusion has been :

use of liquid helium : excluded => cryocooler : mandatory.

3. Analysis of refrigeration needs

What is the objective of refrigeration ?

1) to keep a system at 4.2 K by compensating unavoidable losses due to various causes ; heat radiation, residual gas conduction in an insulating vacuum, conduction by supports.

2) to absorb a certain amount of heat Qg produced in the 4.2 K region by the system during

^ its normal operation.

In small installations, conduction by supports : usually the neck of the cryostat calls ^ for a detailed analysis. When, as is frequently the case, = 0^ conduction in the structure | of a cryocooler, well insulated from thermal radiation, turns out to be a very important

j element of the thermal loss. '

I Let us consider a cryostat with a neck, the cross-section S(x) of which changes in such

a way as to keep a constant thermal gradient between 300 and 4.2 K (practically this section 1 would depart only slightly from the average^one. Hence the following can apply to a real ! cryostat). The heat flux down such a neck Q = k(T) .S(x).dT/dx decreases essentially as k(T) i.e. more or less linearly with T and reaches a value 70 times smaller at 4.2 K where it

is identical to Q^. A regularly decreasing heat flux means that some refrigeration has been

provided all along the neck at each temperature level. I

Without being exactly proportional, the input power requirements and the structural

complexity of a cryocooler are strongly correlated to the minimum work required by a Carnot

cycle work between room temperature and a) 4.2 K concerning Q , b) temperatures distributed between 300 and 4.2 K

at which 69/70 of the heat flux entering through the neck at 300 K are absorbed. The work

related to is then W^ = Q^. (300-4. 2) /4. 2 = 70.4 while that related to the neck

" (l/T)(300-T)dT = 234.5 = 3.33 W^.

Thus, the total required power for the refrigeration of an ideal cryostat reduced to a neck, is more than four times that which is necessary to absorb the residual heat flow which

reaches the 4.2 K level.

In practice cryostat necks display neither a perfectly constant thermal gradient, nor

an exactly linear thermal conductivity. This ratio of 4 is nevertheless a good order of

magnitude at least as a lower limit for real cryostats.

One can wonder why this considerable requirement of refrigerating power is sometimes

unnoticed. There are two reasons : 1) real cryostats are not reduced to necks and the above

ratios 70 and 4 have^to^be reduced by Q total / neck and if some extra power is dissipated at 4.2 K by : (Q^'^Q3)/Q^' 2) finally and more importantly it happens that when one makes

use of liquid helium, the' latter, in excess of its heat of vaporization : 80 j/mole at 4.2 K,

has a specific heat practically constant and equal to the perfect gas value : 21 J/mole.

K

between 4.2 and 300 K, and is thus capable of a total heat absorption of precisely 70 times

the heat of vaporization of liquid helium at 4.2 K. The refrigeration power provided "free

of charge" by the vapor between 4.2 and 300 K is therefore sufficient to absorb nearly all

the heat 70 entering through the neck except for Q which is precisely the source which

generates these vapours.

Practically, taking account of the unavoidable losses of other origin, the neck "costs"

nothing in most of the cryostats which use liquid helium. In cryocoolers the neck will be

resuscitated, then the choice of the refrigeration process should be made preferably among

those offering several cooling levels and even, if possible, a refrigerating power distributed

4
throughout the whole temperature span between 300 and 4.2 K. We will come back later onto

the neck problem after having proposed a number of non-conventional refrigeration processes

the practical realization of which does not seem easy in the actual st^te of t-hp technique

and in the author's imagination. These processes will perhaps be vitiated by trivial error

or by a fundamental flaw but it is hoped that these exposition will help to stimulate

thinking along certain channels which are the present time seem insufficiently explored.

4. C against C p V

The simplest and most immediate way of cooling which is actually the least feasible,

technically speaking, consists of a counter-flow heat exchanger in which a gas cooled at constant volume is subsequently warmed at constant pressure : figure 1 . The gain is Cp - C-^ = R = 8.3 J/mole.K for a perfect gas. The exchanged heat between 300 and 4.2 K is : at constant volume, under an initial pressure of e.g. 3 bars : 3600 J/mole

at constant pressure under 3 bars : 6000 J/mole

under a constant, low pressure (e.g. p < 0.1 bar) : 5970 J/mole

In ( J ) tho. g<u coojU> cut covatant volume., and

In [2] it mAm^ at comtant pKUitixfit.

Thus one can see that in a system operating between these pressures a cooling effect will be

observed as soon as the efficiency attains : (6000-5970) / (5970-3600) = 1.27 % of the theoretical maximum.

Practically any leak or gas friction transfers a part of the gas flux to a Joule-Thomson

cycle which is thermally neutral for a perfect gas and only slightly detrimental for helium

(above 50 K)

.

Although this process has no immediate prospect one can anticipate some technological

advances which would bring it nearer to a practical application : ,s • -, , . •, . ^ • -, • . ^- E (elastic limit) . , ,

. , 1) availability of materials with a ratio -j:;—7 r-p—r-——r- sufficiently high, at room

temperature,

2) the discovery of rubber-like products preserving their elastic properties at low

temperature

,

3) the operation of a system in which the cooling takes place in a two dimensional space

where one can realize constant "volumes" without walls (the surface of a sphere is an example),

4) the introduction of the gas at room temperature under a very high pressure into a matrix from which it could escape only as a superfluid below 2.17 K,

5) if... (any suggestion to fill this 5*'^ paragraph would be welcome).

We will make some supplementary remarks on this process. We will first calculate its refriger5
ation efficiency in the case of a perfect gas. It we take the optimum case, in which the

pressure at the end of the cooling is equal to the constant pressure at which the warming

occurs, one has the elementary compression work at constant temperature Tj(300 K)

dW = -P dV = RTj ^ (for one mole). On the other hand, the cooling effect dQ = (Cp-C^)dT = R dT = RT from which the differential efficiency in the interval ^ (i.e. is do T 0 ^(T1"~T2) "^2 "

dl^ " t7 • average efficiency =

Lo-^^T /T " 17^''^°^ ^l^'^Z compared

to the Camot efficiency n = T2AT1-T2). As an example for Tj = 300 K, T = 4.2 K; n = 0.014;

and on a purely theoretical ground the cycle "Cp vs. C^" becomes more advantageous than the

simple stage Camot refrigerating machine if both of the following assumptions are simultaneously realized :

1) T2 < Tj/2.5 e.g. < 120 K

2) one needs distributed cooling between and and the Carnot machine envisaged

delivers cooling at only one temperature: T2

The "C,3 vs. C " machine has two other peculiarities, when using a perfect gas : a) it provides no cooling at The necessary cooling will have to be delivered by a process

founded on a different cycle. b) it does not "know" if it should produce cold or heat; it

needs to be started in a refrigeration mode otherwise some parasitic heat source will start

it in a warming mode and it will probably finally transform work into heat with an efficiency

1, and this is certainly not a remarkable performance.

5. The gas cocktail adsorption refrigerator

In this system a mixture of gases with different boiling temperatures is introduced under pressure into a column containing a moving adsorbant ribbon with which it moves parallel

while being progressively adsorbed onto the coldest point where the saturation of the

adsorbant is completed by the most volatile gas. The adsorbing ribbon is then led into a return channel under low pressure where the gases desorb successively at temperatures lower

than those at which they have been adsorbed. Thus one produces : 1) a transportation of the

heat of adsorption (practically independent of the temperature) from the lower to the higher

temperature while creating a large AT favourable to thermal exchanges. 2) a counter-flow

cooling of the adsorbed gas of specific heat ==i R by the desorbed low pressure gas of specific

heat = Cp = 5/2 R.

The net effect for a gas mixture is hardly calculable due to the number of parameters

but it is probably very important.

This cycle is akin to a multistage refrigerator using several refrigerating fluids

with distributed boiling temperatures in which:

1) all fluids are incorporated in the same gas flux and use one compressor.

2) two bridges have been established between 50 and 25 K and between 14 and 5 K where mother

nature has not yet produced any body in the liquid state.

What should be underlined is that only a relatively low adsorbant speed is necessary in

order to realize a noteworthy gas flux. As an example, a flow of 1 cm-^ of activated charcoal per second displaces an amount of adsorbed gas of the order of 0.01 mole per second

(at saturation)

.

The heat of adsorption is a function of the degree of saturation of the adsorber. It varies between a maximum (adsorbant degassed) and the heat of vaporization (adsorbant saturated). Table 1 gives approximate values of these heats relative to activated charcoal. [1]

The curves of figure 3 give, for instance, the cooling effect in joules/mole. K in the

following application : adsorbant = activated charcoal, gas = He4, high pressure = 1 bar,

low pressure = 1 millibar. The computation has been made using the formulas [l ] : L = 8600 - 2920 log ^

(V + 112) J/Mole and logj^P = [153 logjQ(V + 112) - 450] . ( 1 /T- .0 16) + 1.9 + logjpV, with P in Torr, V in cm^^^^p/gram of adsorbant, T in K. (RTF = room conditions : 300 K, 1 bar, 1 Mole ^ 24500 cm^) . 6
Pa c Heat of adsorption on activated

charcoal

Heat of vaporization (sublimation)

maximum average

He 2500 1000 80

«2 10000 4000 800

^2 15000 10000 5600

? ? (18400)

Table. J ; App^oxMncutz valuej, 0^ the. hexuU 0^ adJ>oh.ption on acJu.vaX.iid choAcoat [In Joixtt/molz] .

VIqujiz 1 : AmouRi adi^onhoA a {^unction

0^ tmpeAnXuAe..

cuAve A : V = 1 ban. B : P = J mWUban.

C ; dill^^QAmcz beJMdm A and 8

flguAz 3 : Vll^^eAZYvtiaJi heat o^, ad^otption

cufwz V

E

- F

P = 1 boA

P = ? mUlXbaA

global zUzcX = V-E

The curves A and B in figure 2 give the amount adsorbed as a function of T under pressures

of 1 bar and 1 millibar. Curves D and E of fig. 3 give the heats of adsorption (desorption)

per gram of adsorbant and degree K under the same, pressures. Curve F, which is the difference

between D and E, gives the glpbal effect. Finally, curve H (fig. 4) gives the net effect

corrected with the specific heat difference times the difference of the amounts adsorbed : ((A-B) (AV/24500) (5/2R-R)) . Besides this cooling effect which extends from 4.2 to 24 K one

has the available heat of desorption at 4.2 K of 660-520 = 140 cm^/g. of adsorbant, which

will be calculated from the integral curve Q(V) given in figure 5 : Q(660)-Q(520) = 27.5 - 25.8 = 1.7 J/g. of adsorbant which means 63 J/mole circulated.

Thus, the desorbing helium provides cooling essentially in the first zone 5K-14K where

no liquid is available. Hydrogen will provide the bridge between 25 and 50 K and, above the

latter temperature a number of gases exist for going up to room temperature.

7
J/g.of adsorbant.°K J/g.of adsorbant

lJ 30T

F^guAe 4 : coHAHcXzd global. tkojumaJi ^.^ItcJ: CUAV& G : coM.e.cJu.on due. to tkt ctc^fje/ience

-in 6pe.Cyi{i-Lc hmt Cp - C^^

cafwQ. H : coHAzcXed globaZ z^^zciZ = F+G

Vlgyjiz 5 ; Inttgfial htat adiofiption

Wrn' Ml/) dV.

{U In cm^ RTP o;} keLium {^on. 1 ghxm

ad^o^bant] .

6. The Bi-Stirling refrigerator

The machines which operate following a Stirling cycle, the Philips machine for instance,

suffer contradictory requirements,

1) the pistons along which there is a thermal gradient have to move rapidly and offer an important clearance for minimizing the thermal exchanges between parts which periodically

face each other when at different temperatures,

2) regenerators and heat exchangers must present temperature differences and pressure drops

as small as possible and this demands a small gas speed,

3) at very low temperatures the specific heat of metals used in regenerators : copper, lead.,

becomes very small and lead to important volumes hence to important dead spaces.

In conclusion, the Stirling machine which is capable of reaching its full efficiency if

compressions, expansions and exchanges are isodiermal , even with arbitrary dead spaces (reversible cycles) , suffers an efficiency decrease when the dead space increase coincides with

the thermal gradients in the working space of the pistons.

The proposed variant consists of _3.) replacing the pistons with bellows operated slowly

which a) suppresses the shuttle effect and, b) allows isothermal compressions, expansions

and thermal exchanges. The dead space of the bellows reduces to a certain point the specific

power of refrigeration but has no effect on the thermodynamic efficiency, 2) replacing part

or all of the regenerator by a heat exchanger between two machines which are 180° out of

phase. One compressor with a manifold of valves assumes compressions and expansions. A separate system acts on the bellows to operate the gas transfers. Such a system is represented

schematically in fig. 6.

8
lLAAAs======AnAil

Figure 6. Schematic representation of the bi-Stirling refrigerator, in which the compressorexpander (1) and the valves (2, 3, 4, and 5) are operated in synchronism but 90 out of

phase with the bellows (6, 7, 8, and 9) by the mechanical drive (10). The fluid undergoes

a Stirling refrigeration cycle alternately in the left-hand set of bellows and in the righthand set, with displacement through the counterflow heat exchanger (11). Heat is rejected

at the warm (ambient) end through the heat-exchanger (12), and heat is absorbed at the cold

end from the heat-exchanger (13). The outer shell (14) is the vacuum case.

7. The crossed cycles refrigerator

This proposed process combines a) an alternating flow of a caloric fluid in a fixed

column of ideally zero specific heat, with b) a reversible local cycle external to each level

of the column and operating at the same frequency with appropriate dephasing.

At each point of the column the external cycle induces a fluctuation of the temperature

larger than the one of opposite sign, which would be due to the displacemnt of the fluid alone.

A molecule of this fluid follows the cycle represented in figure 7. The line AB would

depict a cycle with zero dephasing. The elliptic curve suggests a real cycle obtained with a given dephasing. The arc NPR is related to the heating part of the external cycle, and the arc RQN to the absorption of heat by the latter.

The heat received by the fluid between Xp and x^j^ and given back to it between x^^^ and Xq

is represented in figure 8.

The overall result is an outgoing fluid at the warmer end at a temperature above room temperature and an outgoing fluid at the colder end at a temperature below T2.

A practical realization would necessarily include a heat capacity effect related

9
T

VIqulHZ 7 : thejmal cycle, a molzcuJLn In VlqaXd S ; kdot txckanQzd In onz cyctz.

the {^leJLd.

to the walls or to the local cycles, which will make this cycle more or less a relative of

the Stirling cycle.

Figure 9a shows a conical piston compressor which, subject to an alternating displacement of small amplitude, cycles the pressure of a gas locally without displacing it. A separate system forces the fluid to oscillate from the warm end to the cold end and conversely.

Figure 9b shows a derived variant in which the displacement system has been omitted

and replaced in its dephasing part by the pressure drop between cylindrical sliding parts.

A net flow will occur if one chooses a displaced cold volume larger than the dead space at the volume warm end.

8. The necks of the cryostats 1

Cryocooler necks require special attention. In order to shed light on this problem

let us consider, at first, the losses in a cryostat neck bathed by the vapours of an evaporated cryogenic liquid. These losses are represented in fig. 10 : Q represent the supplemen- ! tary losses related to the neck^as a function of the losses from other sources (normalization

is made at unity for the value Q^j extrapolated for Qj^ if 9r ^ • '-'^ most cases the thinnest

neck would give a ratio Qj^/Q^ but one can see that Qj^ is practically zero for i Qr/Qn and therefore ° one^obtains a better mechanical sturdiness by increasing the

neck thickness within the ratio Qj^/Qj^ 1'5 and ... one forgets the neck problem.

^

!

IQ
3 C

/ ? —X.

VlguJiz 10 : Lo44e^ tho, necfci compoAtd

to othoA i>0UA.(l2Ji Oj$ £044e6

In the case of cryocoolers one does not have this distributed source of refrigeration

constituted by the vapours heat content. This would result in a heat loss due to the neck

an order of magnitude higher which, in most cases, will overide the other heat loads. If one intends for example to dissipate at 4.2 K all the heat conducted along the neck from

room temperature one will face a disastrous efficiency. Practically all proposed processes

shall include an intermediate refrigeration level and it is more and more evident that those

having the best performance will be 3 staged.

Let us take, for instance, the schematic cryostat of fig. 11 which has 3 levels of

refrigeration at temperatures arbitrarily fixed in geometrical progression i.e. below room n
VZZSa- -d

FXguA.e 7 7 : TypX,caZ cAijoAtat

a : mm md out 300 K

b : l6t 6taQ(L cut 72 K c : 2nd ttaqz out 17.4 K d : 3nd 6tag& cut 4.2 K

temperature (300 K) : 72 K, 17.4 K and 4.2 K. The characteristics of this high performance

cryostat and its calculated losses are given in table 2.

Characteristics

:

useful volume : surface exposed2 liter

to heat radiation : 1000 cm^ ... ^1 emissivity : —

^^2

2 0.01

1 = 3 X 10 = 30 cm neck e

0

0.01 cm

1.5 cm

Zone Q Rad. Q Neck W.Carnot

300-72 K 500 mW 130 mW 2. watt

72-17.4 K 1 .7 mW 13 mW .040 watt

17.4-4.2 K 6. yW 800 yW .003 watt

TabZe. 2 : Typ-LcaJi cAyoitcut ckoAxicXeAAAticii

.

The latter calls for two remarks : 1) The losses due to the neck between the lower

stages are 10 times more important than the radiant heat between 72 K and 17.4 K and 100 times more important between 17.4 K and 4.2 K. This emphasizes the significance of progress in neck

technology and generally in"supports-technology" , that will be demanded by progress in other

12
related fields. 2) Every installation which will make use of liquid nitrogen for cooling of

the first stage is, even with due regard to the scale factor, more than ten times smaller,

ten times lighter, and finally ten times cheaper than a system starting from room temperature.

9. The need of contamination-free compressors

In every refrigerator which makes use of circulated helium, gas purification after

compression, to free it from oil, water and other contaminants such as air or other gases

is a burdensome task which often leads to the need of pfefiodically stopping the system for

clean-up. The use of dry and tight compressors would be a major improvement. But, even with

many potential applications outside the field of cryogenics, it is surprising that a small

contamination-free compressor, is not commercially available.

In fact we have used a two stage rubber diaphragm compressor with slightly modified

heads for a year of uninterrupted service which worked between 0.2 and 3.5 bar abs. Only

a very small contamination has been observed, probably due to diffusion of air through the

diaphragm and the rilsan tubing which leads to monthly warmup and purging of impurities.

Two types of small compressors which should give satisfying results are represented

in figures 12 and 13 . In the first one a very thin metallic bellows compressed by a liquid

properly degassed, compressed itself by another bellows by means of a mechanical device,

offers the advantages of 1) permitting an isothermal compression; 2) being insensitive to thermal expansion of the liquid and hence requires neither pump nor value for the liquid

body and 3) capable of tolerating important depressions on the low pressure side.

VlQiOid 12 : EoUiom c.ompn.u&oK T-jgu/ie. 13 : TheJmocompfLU^oA.

1 . me.cha.nic.cUL boJUiom 1 . hmtyLng mAz

2. tiqiuA 2. mX.cn. on tiquuA viUjiOQcn ^oh. coating.

3. thin boJUiom

The second system is a thermocompressor : a gas volume is warmed by a short ohmic heating of a wire placed in this volume, the resulting increase of pressure opens the exhaust

valve and pushes out a part of the gas. The subsequent cooling induces a depression which

opens the inlet valve and restores the initial conditions. This system has the advantage

of extreme simplicity. It is silent and is readily apt, by adding stages, to produce arbitrarily high compression ratios. It can even be used at very low pressure (= 1 Torr) , subject to the use of actuated valves.

13
References

[1] P. Roubeau, G. deNigohossian, and 0. Avenel, Adsorption de L'helium 4 par le charbon

actif - Colloque international "Vide et Technicens", Grenoble, 1969. Supplement to "Le Vide" No. 143.

14
REFRIGERATION REQUIREMOfTS FOR SUPERCONDUCTING COMPUTERS

R. W. Gueimcy and E. B. FVnt

IBM Thomas J. Watson Research Center

Yoifctown Heights, New Yoik 10598

High speed, low power Josephson junction logic and memory chips densely

packed into a structure immersed in liquid helium may constitute the high performance computer of the future. In this paper we discuss chip cooling, input/output

cable requirements, computer cryosystem design considerations, and refrigeration

requirements.

Conventional gas expansion refrigeration should be suitable for medium to large

size superconducting computers (0.5 to 10 watts). Small sjrstems, such as a 25 mW

future minicomputer, will require innovative refrigerator design that achieves high

reliability at low cost

Key words: Cryogenic equipment; 4K refrigeration; Josephson junction devices; superconducting computers; thermal contact.

1. Perspective

Research is in progress on a wholly new integrated circuit technology that may eventually replace semiconductors in high performance digital equipment [1]. The basic elements are Superconducting QUantum Interference Devices (SQUID) that can be switched from a non-resistive to a resistive state in roughly 10 ps by the magnetic

field from a nearby control line. From these elements, one can build a very high-speed logic family and memory to match [2].

High-speed switches (or gates) alone, however, do not guarantee a high-speed computer. Intercircuit signals, typically travelling at half the speed of light, take 10 ps to go only 1.5mm. In order to avoid excessive signal

propagation times, then, one must pack the circuits (and chips) closely while managing to get power in and the heat out For example, in making an ultra-high performance computer one would want to package the logic and memory

in a cube 9 cm on a side. Such a machine made with present semiconductor technology might dissipate as much as 20kW [2]. The delivery of this amount of power and its removal as heat in such a smaU package would be

impossible with today's technology. A comparable machine made with superconducting circuits might dissipate

roughly lOW [2] which would be easy to remove, even at 4.2K. The power distribution [3] would be via fine superconducting lines. It is the combination of the high speed and the low power of superconducting circuits that may put them in future computers.

An obvious consideration in judging the practicability of a superconducting computer is the refrigerator, and so

there is interest in the state of this technology and in what improvement in reliability and reductions in cost could come in the next ten years. We present in this paper, as best we can predict, what a superconducting computer's

cryosystem requirements might be. Let us take a more detailed look at these by starting at the heat source.

2. Heat transport: circuit to sink

Most of the heat generated by the circuits travels through several thin film layers to the chip substrate (a 6.4 mm square of silicon) which is isothermal because of its high thermal conductivity (~ 2 W/cmK). Most of this heat

15
then passes out the non-circuit side of the chip to a refrigerated heat sink such as a Kquid helium bath. Heat

transport in the bath is by natural convection, and heat transfer at the chip surface is aided by the stirring action of

local boiling.

Figure 1 shows the expected chip temperature rise as a function of the chip power [4], and also shows

anticipated power levels for logic, fast memory (cache), and main memory chips. These are essentially the same

whether or not computations are taking place. The somewhat poorer heat transfer obtained as the chip power is brought up initially probably is due to incompletely developed nucleate boiling. Note that the substrate temperature

for the three kinds of chips should run from 7 mK to 240 mK higher than the bath temperature. Note also that the highest chip power is only about 4% of the nucleate boiling limit, and if this power were increased tenfold then the chip temperature would be less than 400 mK above the bath.

In a small computer, one might choose to dispense vrith the bath and make a greased thermal contact to the

chips. At 4.2K, the grease to silicon boundary resistance is negligible compared to the resistance of practical thicknesses of grease. The heat transfer characteristic for a 2/zm thick, 6.4 mm square of silicone grease [5] is shown

in Fig. 1.

For the interferometers currently in use at IBM, the maximum desirable chip substrate temperature is 4.5K,

and one can expect successful operation of a machine- whose chips vary in temperature from 4.0 to 4.5K. Then, for example, one could operate with a total spacial and temporal temperature variation from 4.0K to 4.2K, and so the

variously loaded chips could run up to 0.3K warmer.

100

£ 10

a:Ld

o

0.

0.1

—LOGICI I I I I M|

T=4.2K

T—I I 1 1 1 1 1| 1

I 1 1 1 1 m

END OF NUCLEATE

BOILING

—FASTMEMORY

CHIP GREASED—/

TO SILICON

(EST) y fl

0

I6

/ /

CHIP IMMERSED IN

LIQUID HELIUM

MAIN MEMORY

lo^LUJJji J I i I I I I ll J ' I I I I n

10 100

CHIP TEMPERATURE RISE (mK)

1000

Figure 1: Chip Cooling Characteristics

16
3. Medium to large computer cryosystem

The logic and memory chips of a medium to large computer might be mounted on cards that plug into a box-like

structure as shown in Fig. 2. This would be immersed in liquid helium, and the chips would be cooled by natural convection of the liquid up the vertical channels. A large computer might require a box structure 9 cm on a side and

I would dissipate approximately lOW of heat into the bath. A medium capacity computer might dissipate 0.5W to 2W.

i Communication to room temperature support equipment would be via a multiple stripline input/ouput cable as shown in Figs. 2 and 3. Each copper line and associated groundplane would have a cross-sectional area of roughly

6.1 X 10"^ cm2, and an ultra-high performance computer might have 400 lines. If these were thermally grounded at 75K and 18K as shown in Fig. 3, with a 20 cm length between stages, then the heat leak to these grounds would be

1.2W and LOW, respectively, and the heat leak to the 4.2K bath would be 0.3W which is small compared to the computer's lOW. Perhaps 20 of these lines would supply the power and so would have noticeable resistive heating

that could bring the total I/O heat loads for the large system to 4.7W, 2.0W and 0.8W at 75K, 18K and 4.2K,

respectively. One sees that the I/O lines will have little effect on the 4.2K refrigeration requirement

Liquid helium

Interboard

flexible connection

Figure 2: A Superconducting Computer Package Unit

17
EoOCO

18
I

I There are three principal functions that we wish to achieve in a medium to large sized computer cryosj^em:

(1) Uninterrupted computer operation during refrigerator semcing,

(2) Uninterrupted refrigeration operation during computer removal, and

(3) A semi-annual or longer service interval An integral refrigerator/dewar cryosystem that should meet these requirements is shown in Fig. 3. If the

I refrigerator should fail or be shut down, then the thermal couplings (TDl, TD2, and condenser) isolate it from the

I bath which then provides temporary refrigeration by gradually boiling away. When the refrigerator is returned to

j operation, helium gas is liquefied to replenish the bath. The computer can he removed by pulling it up into a special

' room temperature service housing (not shown) that serves as an air lock. Heat from the I/O cable and from

radiation and conduction in the dewar neck is intercepted by 75K and 18K heat stations that are part of the computer support structure. Each of these delivers its heat across a narrow annular gap to a refrigerated ring in the dewar neck waU. The compressor, refrigerator, and thermal couplings as shown schematically in the figure have been designed and built for IBM by the Advanced Products Division of Air Products and Chemicals. Further

discussion of the design and performance is given by R. Longsworth and E. Flint et al elsewhere in these

{ proceedings.

An important feature of the refrigerator/dewar interface scheme is the simplicity of the connection between the

refrigerator and the couplings. This consists of six small lines that carry the JT gas at various temperatures, and

these can be made flexible (for vibraton isolation) or longer (for remote refrgerator placement). For example, in order to simplify dewar design or to make the system fit in tight quarters, one might place the refrigerator in a

separate enclosure joined to the dewar by an evacuated link that carries the interconnecting tubing.

Good magnetic shielding is requried for the logic and memory devices. This can be provided by ferromagnetic

shields and/or a superconducting shield as shown in Fig. 3. Materials having low remanent magnetization such as

silicon, glass, polymide, aluminum, selected beryllium and copper, and molybdenum must be used for structures within the shield.

4. Cryogenic challenges

Refrigerator, interface, and dewar technology suitable for commercial medium to large superconducting

computers is available, and we look for continued improvement in reliability, efficiency, and cost. Obtaining high

reliability is of greatest importance. Installations in aircraft or satellites will pose obvious additional problems that challenge present day cryogenic technology.

The commercial viability of smaU superconducting computers (power < 0.5W) depends on both the reliability and the cost of the cryogenic system. In the extreme case of a future minicomputer, one would like to have a 25 to 50 mW 4K refrigerator for a few thousand dollars. This caEs for real engineering innovation. One could think of simply using a large dewar, but a reasonably sized system is likely to need monthly fillings and the logistics of this for a machine destined for a diverse market look unattractive. A schematic drawing of a hypothetical future minicomputer is given in Fig. 4. There would be no liquid helium

bath and the chips would be cooled by greased contact to a silicon heat sink. The refrigerator would be capable of

rapid cooldown (~ 10 min with the computer off). It might be a gas cycle machine depending on an external compressor, or it might be an innovative self-contained system where only exhaust heat and electrical or mechanical

power pass through the interface. While magnetic, para-electric, and other cooling techniques are possible, gas expansion may be the most practical. An isothermal expansion of 13 mg of helium gas at 4K from 1 cm' to 2 cm'

in one second can, in principle, produce 80 mW of refrigeration. The challenge is in delivering the expansion work

out of the system and in recycling the gas.

5. Summary

Superconducting computers requiring closed cycle 4K cryosystems of up to lOW capacity may be on the market

in the future. The medium to large size machines will be cooled by convection in a refrigerated liquid helium bath. Current refrigerator, interface, and dewar technologies are adequate for this application, although improved

reliability, efficiency, and cost would be desirable. The future commercial viability of small superconducting

computers depends on the development of new low capacity, low cost, but highly rebable 4K refrigerators.

19
30 to 60 cm

REF. I/O

REFRIGERATOR

1st STAGE

COMPUTER I/O

-MAGNETIC

SHIELDS

-I/O CABLE

-CABLE CONNECTOR

-THERMAL

BASE FOR

CHIPS (SILICON)

25 TO 50 mW

OF REFRIGERATION NEEDED AT

3rd STAGE

-VACUUM a

RADIATION SHIELDS

-10 to 20 cmFigure 4: A Hypothetical Superconducting "Future

Minicomputer". TG = thermal ground

6. References

[1] Matisoo, J., The Superconducting Computer, Sci. Am. 242, [5], 50-65 (May 1980).

[2] Anacker, W. et Josephson Computer Technology, IBM J. Res. Dev., 24, [2], 107-264 (March 1980).

[3] Arnett, P. C and Herrell, D. J., Power Design for Gigabit Josephson Logic Systems, IEEE Trans. Microwave

Theory Tech., MTT-28, [5], 500-508 (May 1980).

[4] Flint, E. B., Jenkins, L. C, and Guernsey, R. W., to be published.

[5] Kreitman, M. M. and (Callahan, J. T., Low Temperature Thermal Conductivity of Several Greases, Rev. ScL

Instrum., 40, 1562-1565 (1969).

20
DEVELOPMENT APPROACHES FOR LONG-LIFE CRYO-COOLERS

Ronald White

William Haskin

Flight Dynamics Laboratory

Air Force Wright Aeronautical Laboratories

Wright-Patterson Air Force Base, Ohio 45433

Developmient of cryogenic refrigerators ( cryo-coolers ) capable of

operating for over three years without maintenance is being sponsored

by the Flight Dynamics Laboratory for several Air Force missions.

Future cryo-coolers for superconducting devices will need similar high

reliability and durability. Therefore, the general approaches used and

the lessons still being learned will be described as they apply to all

long-life cryo-coolers.

Highly reliable and durable cryo-coolers are developed in a carefully

planned and controlled process. Required operating conditions will

influence the selection of the basic type of cooler, but the selected

concept must have an inherent long-life potential. The principal longlife cryo-coolers being developed by the Flight Dynamics Laboratory

are the Vuilleumier (VM) , turbo-Bray ton and rotary reciprocating

refrigerators. These machines either use gas film bearings to avoid

rubbing contact or have lightly loaded bearings and seals with low wear rates.

Careful development is necessary to realize the long-life potential

of the coolers. This includes detailed analysis of stress levels for

all parts of the cooler, adequate quality controls, contamination

control, and reliability analysis. Final verification of adequate

durability, with acceptable cooler performance, is demonstrated by

reliability growth in a test, analyze, and correct program.

Key words: Component development; contamination control; cryogenics;

helium retention; long-life; refrigerators; reliability; rotaryreciprocating; turbo-Bray ton ; Vuilleumier.

1. Introduction

Cryogenic refrigerators (cryo-coolers) capable of operating for over three years without maintenance are being developed by the Flight Dynamics Laboratory for several Air Force

missions. Other applications may permit occassional maintenance, but good reliability and

dependability will always be essential. This is especially true for future superconducting

devices. The experience gained and the lessons still being learned in the development of

long-life cryo-coolers will benefit future programs.

Active development of several types of long-life cryo-coolers is in progress. Different

design approaches are being pursued until sufficient success is obtained to satisfy various

requirements. Air Force efforts include Vuilleumier (VM) cycle, turbo-Bray ton and rotaryreciprocating (r3) Brayton cycle refrigerators. These coolers produce temperatures as low

as IQOK with cooling capacities ranging from O.lW at 10°K to over lOW at temperatures

between 70° and 100°K. The turbo-Brayton and rotary-reciprocating refrigerators can be

designed to produce temperatures down to 40K or lower, and component development is planned

for improved refrigerators to provide cooling at 4°K. The cryo-coolers presently being

developed will be used to describe the approaches for obtaining long-life.

21
The VM cooler achieves internal pressure changes by gas displacement to different

temperature volumes rather than by mechanical compression. The total gas volume in the

cooler remains constant. Forces on bearings and seals are therefore sufficiently low to allow long operating times. Heated cylinders with dome s maintained at 960^K are used for

thermal compression, and gas expansion in the cold cylinder provides refrigeration. A

small electric motor is used to overcome friction and control the motion of the displacers

in the cooler. Thermal regenerators are used to exchange heat between incoming and exiting

gas streams, to obtain adequate efficiency, for this reversing flow refrigeration cycle in

much the same way as counterflow heat exchangers do in continuous flow refrigeration cycles.

Figure 1 shows a schematic of a VM cooler which has two hot cylinders and three stages of

expansion. This cooler [l] is being developed at the Hughes Aircraft Company for long

duration Air Force missions.

The turbo-Brayton cooler uses small, high speed turbine wheels supported on gas film

bearings to achieve compression and expansion of the working gas in the machine. Since

there are no rubbing parts, the cooler should be capable of long term operation. A two- stage development model turbo-Brayton cooler [2] is being assembled at the AiResearch

Manufacturing Company of California. Figure 2 shows a schematic for the arrangements of this

cooler

.

The concept of combining the efficiency of recpirocating refrigerators with the longlife capability of components suspended on gas film bearings provided the basic approach

for rotary-reciprocating refrigerators (R^). This type of machine is characterized by

pistons rotating on their central axes at 1200-1400 RPM as they reciprocate at a cycle rate

twice their rotational speed. This motion provides porting action to control the flow of

the working gas which is compressed and expanded in different sections of the machine to produce refrigeration. The pistons are driven by electromagnetic actuators. Counterflow

heat exchangers are used to implement the reverse Brayton cycle. A two-stage unit of this

type has been tested and components are being developed by Arthur D. Little, Inc. [3] for

a three-stage unit. Figure 3 shows the basic arrangement for a R-^ system.

The Stirling cycle is often used for comparison because it is potentially the most

efficient. It is a refrigerative cycle with a cold section very similar to the VM cooler.

However, the gas is compressed by a mechanical piston instead of heated cylinders. The

forces on the bearings and piston seals are therefore significantly higher than the VM

which has resulted in wear and shorter life. Recently NASA contracted with Philips Laboratories[4] to design and build a long life linearly driven Stirling cooler utilizing non-contacting magnetic bearings and clearance seals. This cooler will be described in another

paper at this conference.

As indicated above, there are reasons to expect each type of cooler to be capable of

operating for long periods of time. However, careful development is necessary to determine

the true long-life potential. Methods used in these developments are described in the

next section.

2. Development approach

In planning the development of a long-life cryo-cooler, the first essential step is to choose a type of machine which has an inherent capability to operate for long periods of

time under expected environmental conditions. Although this is obvious, the applicable

technology may need to be examined in some detail before making a decision. For example,

an oil lubricated reciprocating compressor is generally accepted as having better durability and reliability than a dry lubricated reciprocating compressor. However, use of

the oil lubricated compressor depends on the ability to separate the oil from the working

gas in the refrigerator. Periodic replacement of the oil separation filters will be

necessary. Thus, the oil lubricated compressor is not acceptable for use in a location

where equipment cannot be- maintained . The maintenance problem might be circumvented by

use of devices such as rolling diaphram seals, but these add further complications which

must be considered. These include a full-time oil support for the diaphram seals, a system to dispose of the helium that permeates into the oil through the diaphram seals, a helium makeup and control system, and devices to compensate for oil displacement and

expansion.

22
23
HEAT

TRANSPORT

FLUID

INPUT c^__

POWER O", I

COMPRESSOR

SUBSYSTEM

AC-AFTERCOOLER

C-COMPRESSOR STAGE

M-MOTOR

R-RECUPERATOR

A-ALTERNATOR

T-TURBINE

SECONDARY

COOLING LOAD

INTERFACE

MULTILAYER

INSULATION

COLD SECTION

PRIMARY

> COOLING LOAD

INTERFACE

FIGURE 2. TURBO-REFRIGERATOR SYSTEM SCHEMATIC

24
25
other factors which influence the choice of a basic concept for a long-life cryo-cooler

include power consumption and weight restraints which may be imposed. Turbine machinery

with gas film bearings should operate for very long periods of time, but the efficiency is low unless a high cooling capacity is needed. The rotary-reciprocating refrigerator has a potential for very long-life and high efficiency but with higher weight and more complicated

control circuits than other coolers. The VM cooler does not have as great a life potential

as the gas bearing coolers, but it will have significant life because of the light loads.

Its efficiency will not be as high as the R^, but its weight is considerably less, the

controls are simple, and the design more mature. Much better efficiency should be obtainable with Stirling cycle machines if novel means can be implemented to achieve long-life.

The relative importance of different factors must be evaluated in making a decision.

Cryogenic refrigerators being developed by the Flight Dynamics Laboratory for long term

continuous operation generally use gas film bearings to avoid rubbing contact or use lightly

loaded bearings and seals with low wear rates. The turbo-Brayton and rotary-reciprocating

refrigerators with gas film bearings may be subject to long term gas erosion and contamination effects, but their good durability potential is generally recognized. Durability of

VM coolers may be somewhat questionable, but thermal compression and the constant volume

aspect of the VM cycle provide significant benefits. Bearing loads and operating speeds

can be an order of magnitude lower than for conventional Stirling coolers with equivalent

capacity. The low forces permit long-life to be achieved with dry lubricated bearings.

A low pressure drop across the sliding seals allows innovative, long lasting seal designs.

Careful material selection along with thorough component specifications to assure adequate

quality are leading to significant progress in extending the demonstrated life of VM

coolers

.

Other concepts for long-life coolers are being pursued by other organizations in

relation to their particular requirements. NBS has demonstrated a low speed, low capacity

Stirling cycle cooler [5] that is well suited to laboratory or similar use. NASA is sponsoring the development of a Stirling cycle cooler [4] which employs magnetic bearings

to avoid rubbing contact. The combined efforts of all organizations should provide the

technology to increase the potential for long life coolers in the next five years.

2.1 Component development

After the basic cryo-cooler concept is chosen, the long-life potential can be realized

only by careful analysis and controlled fabrication of each part or component. Sufficient

analyses must be performed to determine the mechanical and thermal stress levels for each

part. The fatigue or wear limits for each part must then provide a margin of safety when

compared to the durability requirements of the cooler. Those parts which may cause failures

are identified for special emphasis in the development phase.

During the development of VM coolers, many early problems, such as heater life and

mechanical interference forces, were eliminated by bench testing and analysis. A special

material had to be found for the rider rings which guide the hot end of the displacers in

the hot cylinders. Experience with sliding seals in the coolers indicated that simply

finding a better material would not provide adequate life. Therefore, new designs are being evaluated along with new materials and better control of the surfaces on which the

seals rub. These are examples of the detailed effort which is now showing a payoff for

the total cooler.

In the case of the rotary-reciprocating refrigerator, technical issues include the

design and fabrication of gas film bearings, the complexities of electrical drive circuits,

and the fabrication of high effectiveness heat exchangers. Contamination control is also

very important in relation to materials selection, assembly methods, and gas management.

These problems have been identified, and solutions to them are being verified as development continues. Since the electrical control circuits for the rotary-reciprocating refrigerator are rather complicated, special emphasis has been placed on the basic design of

these circuits to provide highly derated part operation with well established functions.

Electrical problems with other coolers have shown that reliability and durability problems

are not limited to mechanical devices. Technical issues for the turbo refrigerator include

design and fabrication of compliant gas bearings, design and fabrication of high effectiveness heat exchangers, development of a high speed motor and controller, and the design and

26
fabrication of small turbine wheels. Quality control and contamination have been given a high level of attention. Included also, is a continuing effort to increase the efficiency

of individual components such as heat exchanger effectiveness, compressor adiabatic efficiency, turbine adiabatic efficiency, and reduction of seal leakage and bearing losses.

2.2 Quality assurance

An adequate quality assurance program must be implemented during the development of a cryo-cooler. This program provides for verified identification of materials and complete

inspection of all purchased and fabricated parts. This information is necessary to assure

a reasonable probability of successful cooler operation and to help analyze the cause of

difficulties which do occur. Part failures have been caused by materials which did not actually contain a specified constituent. In a similar manner, machined parts must be

inspected for proper dimensions and quality of workmanship. Component assembly procedures

must be reviewed for adequate process control, and the components should be inspected to assure that specified steps, such as heat treatments, were actually accomplished. Acceptance of this extra quality control burden is essential to the final success of a long-life

cryo-cooler

.

2.3 Contamination control

One of the most significant problems in developing long-life coolers is the elimination and control of contamination. The problem is usually broken into two categories:

particulate contamination and gaseous contamination. Particulate contamination can result

from inadequate cleaning, manufacturing defects, flaking of coatings and wear debris from

bearings, riders, and seals. The effect of particulate contaminates varies from the

scratching of highly finished close tolerance bearing and seal surfaces to the partial or total plugging of heat exchangers, regenerators, and other small gas passages. Scratching

of surfaces can result in accelerated seal wear, bearing failures, additional particle

generation, and the jamming of the refrigerator mechanism.

Advanced design concepts, cleanliness standards, and particle traps are utilized to eliminate or reduce particulate contamination. Design techniques include the elimination

of rubbing parts by the utilization of gas bearings and non-contacting seals. Examples of

this technique are the turbo-Bray ton and rotary-reciprocating refrigerators.

In coolers that do have rubbing parts, the quantity of particulate contamination can

be reduced by reducing the load on the parts and by dry lubrication. An example of this

is the Vuilleumier refrigerator which utilizes heat instead of mechanical energy as the

primary input power. Inside the refrigerator, the pressure drops across the thermal

regenerators are only a few pounds per square inch. This results in low loads on the

dynamic (sliding) seals, displacer riders, and bearings which, coupled with their slow

speed, results in low wear, low contaminate production, and long-life.

Other techniques to reduce particulate contamination are to keep trapped dirt sources, such as motor windings, out of the working fluid (helium) or to hermetically seal the con- tamination source. Successful techniques have been to place the pressure wall between the

solid metal motor rotor (inside the cooler) and the motor stator with its windings outside

the cooler or to "can" the motor stator.

Cleaning prior to and during assembly is required for all cryogenic coolers but is especially critical for the extremely close clearance gas bearing refrigerators. Clearances

in these bearings are typically measured in millionths of an inch.

Once assembled, the continuous flow coolers utilize filters to trap particulate con- tamination missed by cleaning or generated within the cooler. Reversing flow refrigerators

such as the VM and Stirling utilize thermal regenerators which, because of the reversing

flow, tend to be somewhat self-cleaning. Particulate traps can be created by adding small,

deep groves perpendicular to the working fluid flow that allow contaminate or wear particulates to drop in and become trapped. The location and number of these type of particle

traps must be carefully considered since they decrease refrigerator performance by increasing refrigerator dead volume and pressure drop.

27
Condensation or freezing of gaseous contamination is an even more severe problem than particulate contamination in cryogenic coolers since, at the very low temperature that some of

these coolers produce, everything except helium is a liquid or a solid. Gaseous contamination can clog Joule-Thomson valves, clog heat exchangers, partially clog thermal regenerators, create thermal shorts from cryogenic regions to warmer regions of the refrigerator,

and possibly even stall the refrigerator mechanism. Sources of gaseous contamination

include organic materials such as seals, riders, bonding agents, plastic parts, motor

wire insulation, and the gaseous contaminates adhering to the surfaces of the metal parts

during assembly. Gaseous contamination has also been obtained from untested dirty helium

fill bottles. Due to the difference in partial pressures of the contaminates, between the

pure helium inside the refrigerator and the atmosphere outside, refrigerators with large

static seal areas may experience additional long-term gaseous contamination problems

caused by permeation of contaminates past the seals. Design techniques to reduce gaseous

contamination include: elimination of organics within the refrigerator, hermetically sealing the motor windings or placing the windings outside the working fluid, and the use of

welded joints whenever possible. Operational techniques include testing, certifying, or cold trapping the helium fill gas, careful purging of the cooler after assembly and after

break-in, and bakeout of the cooler parts and the assembled cooler, which is limited by the

high temperature tolerance of the seals and epoxy or solder bonding agents.

2.5 Helium retention

All of the refrigerators of interest utilize gaseous helium as the working fluid. Since

helium molecules are small single atoms, it is one of the most difficult gases to contain

for long periods of time. Techniques to maintain adequate working fluid within the refrigerator are helium makeup and helium retention. Helium makeup has been confined to large

refrigerators and utilizes a high pressure bottle, pressure reducing regulator, valves,

pressure sensors, and control logic. This adds parts to the refrigerator which in turn can affect system reliability. Helium retention has been accomplished by various means. For

short life coolers, single or double "0" rings have been used. Since rubber "0" rings are permeable to helium, longer life coolers may have an "0" ring to withstand the internal

refrigerator pressure fluctuations backed by a soft metal cut ring seal or indium coated

metal "K" seal for positive sealing. These seals require careful preparation of the sealing surfaces and replacement of the seal after the joint is opened. The best way to seal

joints is by welding. This can be considered after the refrigerator development phase has

been completed and there is no longer a need to frequently disassemble the refrigerator.

Clever weld flange designs allow the weld to be ground off and opened and then rewelded

one or more times without destroying the expensive parts.

The detection and location of very small leaks is always a challenge especially on refrigerators of complex geometry. Leak isolation is currently being investigated under

the VM effort.

2.6 Reliability

Reliability can be enhanced if the refrigerator designer pays careful attention to the

environments and conditions that the cooler will experience during developmental testing

and its ultimate use. Design simplicity with attendant low parts count enhances reliability. Since long-life coolers are not high production rate items, 100% inspection is

possible and appropriate. Inspection records of wearing parts can be used to establish wear rates that may be extrapolated to indicate refrigerator life. The use of flight

qualified electronic parts, although more expensive, appears to prevent numerous breakdowns later during refrigerator life testing or use.

In selecting the type of cooler to be developed, it is sometimes necessary to choose

between a cooler that has good reliability associated with simplicity and a low number of

parts but is limited by wearout of parts as compared to a more complicated machine that

has lower reliability but better durability. Awareness of this conflict in relation to operational requirements can aid in selecting the best cryo-cooler for a particular

application.

The quality of the support equipment should also be carefully considered before acquisition. Cooling cart (for heat rejection) breakdowns have been a continuing problem that

28
interrupts cooler testing and causes delays. Power supply and instrumentation malfunctions

have resulted in many lost hours of operation and have confused the malfunction analyses.

Adequate spares should be acquired to prevent long term breakdowns of the entire system.

Reliability and durability can be verified in several ways. Current Flight Dynamics

Laboratory programs use the Test, Analyze, and Correct (TAG) method to demonstrate the

life and increase the reliability of long-life refrigerators. This method provides

reliability information on the failure modes and mechanisms of critical components while

undergoing development and life testing. Reliability improvement (growth) results when

failure modes and mechanisms are discovered, identified, and their recurrence prevented

through implementation of corrective action.

3. Conclusion

In summary, working across all refrigerator technical fronts, there is a high expectation that the technology for five year coolers will be available within the next five years.

Development of a highly reliable, long-life cryo-cooler must be a carefully planned and con- trolled process. Required operating conditions will influence the selection of the basic

type of cooler, but the selected concept must have an inherent long-life potential. This

potential is then realized through careful analysis of stress levels for each part and

adequate quality control. Contamination control must also be implemented. Good reliability

can be obtained by limiting the number of parts in a relatively simple cooler and by assuring adequate design margins for all parts. Final verification of adequate durability with

acceptable cooler performance is demonstrated by reliability growth in a test, analyze, and

correct program.

A. References

[l] Doody, R. D. , The High Capacity Spaceborne Vuilleumier Refrigerator, Paper No. 2A5-15

presented at the SPIE Annual Technical Symposium, July 1980.

[2] Wapato, P. G. and Norman, R. H. , Long Duration Cryogenic Cooling with the Reversed

Bray ton Turbo-Refrigerator, Paper No. 245-17 presented at the SPIE Annual Technical

Symposium, July 1980.

[3] Breckenridge , R. W. , Refrigerators for Cooling Spaceborne Sensors, Paper No. 245-16

presented at the SPIE Annual Symposium, July 1980.

[a] Sherman, A., Gasser, M. , Goldowski, M. , Benson, G. , and McCormick, J., Progress on the

Development of a 3-5 Year Lifetime Stirling Cycle Refrigerator for Space, Paper No. KC-1 presented at the Cryogenic Engineering Conference, August 1979.

/ [5] Sullivan, D. B. , and Zimmerman, J. E. , Very Low-Power Stirling Cryocoolers using

Plastic and Composite Materials, International Journal of Refrigeration, Volume 2, No. 6 pp 211-213, November 1979.

29
THEORETICAL ANALYSIS OF A 3-STAGE STIRLING CYCLE CRYOCOOLER

Stuart B. Horn Mark S. Asher

Far Infrared Team, Systems Integration Division

Night Vision & Electro-Optics Laboratory

Fort Belvoir, Virginia

Previous analyses of three stage coolers have neglected

the real gas equation of state of helium. Those analyses have

therefore neglected the compressibility effects that occur

below 20K. This paper derives the thermodynamic performance

of three stage Stirling type coolers using the real gas equation

of state. An analysis of a small, efficient three stage cooler

is given with the low temperature gas effects illustrated.

Key words: Cryocooler; cryogenic; helium; low temperature;

refrigerator; Stirling cycle.

The three stage Stirling Cryocooler (Fig. I) consists of a motor driven reciprocating

helium compressor and a cold finger containing three expansion chambers separated by

porous regenerators. The displacer-regenerator assembly may be driven by a cam or

slider crank mounted on the same shaft as the compressor motor. Alternatively, a

linear solenoidal drive may be used to drive the displacer assembly with no side thrust. The motion of the displacer could be easily changed with the solenoidal drive whereas

a new lobe would have to be ground if a cam were used to drive the displacer.

Cooling is produced in the three expansion chambers by phasing the displacer motion

with the pressure pulse generated by the compressor so that the working gas expands

at high pressure and contracts at low pressure (Fig. 2). The regenerator? serve to maintain the temperature 'difference between the three expansion spaces in the cold

finger. As gas travels down toward the cold end of the cold finger it is precooled by

the porous regenerator matrices. The matrices are recooled by the gas after expansion as it flows back to the ambient end. The matrix materials should have high

volumetric specific heats relative to the compressed helium. If this is so, the axial

temperature distribution In the regenerator matrices will be essentially constant in time and linear between the expansion spaces. Because the heat capacity of

the solid drops so severely at low temperature (below 30K) and the helium

heat capacity is pressure dependent in the same range, this assumption is

clearly not exact at the lower end of the machine.

In most cases, refrigeration will be required only at the cold end of the cold finger.

The upper two stages should then be designed to produce no net refrigeration. They

should serve only to eliminate the flow losses associated with the regenerators.

The computer program used to predict the performance of the cryocooler is a modified

version of the one described by Horn &. Lumpkin (1) of the Army Night Vision

Laboratories. This is a second-order isothermal analysis with decoupled losses. The

original program was modified to include a three stage cold finger and real gas

equations of state in the cold end. The program contains an iterative procedure to find the correct mass distribution in the machine and in iterative solution to the real

30
VREG1

(VOID VOLUME)

VC1. T1

VC2. 12

VREG2

VC3. T3 VZ, TAMB

VREG3

[ ] VDLINE

VAMB, TAMB-

\

VCR, TCR

FIG. 1. THREE STAGE STIRLING CRYOCOOLER

31
32
( START

j

/ READ DATA /

I

INITIALIZE MASS DISTRIBUTION

BETWEEN WORKING AND

CRANKCASE VOLUMES

INCREMENT TIME

AND FIND NEW

DISPLACER AND

COMPRESSOR

POSITIONS

INITIALIZE DISPLACER AND

COMPRESSOR POSITIONS

FOR t = 0

SOLVE FOR THE SYSTEM

PRESSURE FOR A PARTICULAR

DISPLACER AND COMPRESSOR

POSITION (ITERATIVE SOLUTION)

ADJUST MASS

FRACTION

FIND GROSS COOLING

AND SUBTRACT

LOSSES

OUTPUT RESULTS

END

FIG. 3. PROGRAM FLOWCHART

33
gas equation of state. The non-zero Joule-Thomson coefficeint has not been

completely included in this model.

A thermodynamic model of the cryocooler is shown in Fig. 1. In order for there to be

no long term leakage of gas around the compressor piston seals, the time averaged

pressure in the crankcase must be equal to the cycle averaged pressure in the working

volumes:

Per dt = i^-J dt (1)

To satisfy this condition the program solves these integrals numerically and adjusts the

amount of mass present in the working and crankcase volume. In order to evaluate the

cycle averaged pressure the program solves the real gas equation of state numerically

for every time step. There are then three nested iterative procedures contained in the

program as shown in the flow chart (Fig. 3).

First, a rough estimate must be made of the mass distribution in the machine. The

total amount of mass that should be present in the entire working volume can be found

by summing the masses in the separate subvolumes:

Mw - f£ \ M + 2 VregZ . M + M . 2 Vreq3 , Vdline

""l " R I JZ T2+T3 T3 Tl T3+Tamb Tamb

Vamb Vz 2 Vregl "1 Tamb Tamb T1+T2 J

It has been assumed that the machine begins its cycle with both the compressor and

the dispiacer at mid-stroke. Also, all isothermal subvolumes are assumed to already

be at their equilibrium temperatures. The subscript 1 refers to the first mass

distribution iteration. The mass in the crankcase can be found from the following

relation: „ Pc Vcr

^'^''l = R- T^ (3)

Where the volume of the crankcase is assumed to be essentially constant. The total mass of the system is then:

M^Q^ = Mw^ + Mcrj (4)

We now define the first estimate of the working gas mass fraction as:

"tot

In order to proceed further we must find the pressure in the working space as a function of time for one complete cycle. During this first cycle we will assume that

there is no leakage of mass from the compression space into the crankcase. For some

general point in the cycle, then, we can relate the instantaneous pressure in the

working space to the mass fraction by again summing up the mass contributions of the

subvolumes:

Mtot Fi = ^i f Mi +

2 Vreg2 ^ Vc3 ^ 2 Vreg3 ^ Vdline moz i-j pi

T2 ^ T2+T3 T3 ^ T3+Tamb ^ Tamb

^ " T^i^bO ^ P'-^Sl Vregl + pel Vcl.

(6)

34
Where the following assumptions have been made:

1) All subvolumes are isothermal.

2) All subvolumes in the working space are in pressure equilibrium (no pressure

drops).

3) The gas in the regenerators is at the arithmetic mean temperature of the

expansion spaces that they connect. This is a simplifying assumption.

While the temperature is a linear function of position, the gas density

varies logarithmically with z.

i^) The ideal gas law is correct in all subvolumes except the cold expansion

space and the cold regenerator.

The subscript i in Eq. (6) refers to a particular instant in the cycle while the subscript j refers to a mass fraction iteration. The motion of the compressor and the displacer can be either sinusoidal or matched constant acceleration parabolic sections. The

displacer and compressor motions define the instantaneous volumes Vcl., Vc2., Vc3-,

Vamb, and Vzj. Solving Eq. (6) for the system pressure we have: ^

^

^

Pwi =[Mto t Fj - (preglj Vregl + pcl-j Vclj R Denom

where:

Denom- Vc2i ^ 2Vreg2 ^

VcSj ^ 2 Vreg3 ^ Vdline + Vamb +

Vzi ~

T2+T3 13 T3+Tamb Tamb Tamb Tamb

All terms in Eq. (7) are known except for pregl; and pel.. An iterative procedure

must be used to find a pel. and a pregl. that result in pressure equilibrium throughout

the machine. The flowchart in Fig. 'fa sViows how this procedure works.

First, a rough estimate of the instantaneous pressure in the machine is made by using

the ideal gas law to find pel and p regl in Eq. (7):

[fj Mtot -f 2Pwi Vregl + Pwi VcliTl R I VR(T1+T2) m ]\ Deno

Pwi = \Fj Mtot -f2Pwi Vregl + Pwi Vcli^^ j R (8)

Denom

Solving for Pwj we have:

PWi (guess) = R Mtot Fj [ 1 + 1 ( 2 Vregl ^ Vc^iM , . Denom L Denom \ Tl + T2 T2 /J ^

'

35
GUESS SYSTEM PRESSURE

USING IDEAL GAS LAW:

RMtotF . P (GUESS) = ^I DENOM 1 + DENOM

1 /2Vregi , Yci\y

siomVt, + T2 tJJ

GUESS COLD EXPANSION SPACE

DENSITY:

0.2P..(GUESS)

, (GUESS) =

RT, T

CALCULATE INCREMENT FOR Q^^ :

Ae=0.28''w<.5^

RT,

SOLVE FOR SYSTEM PRESSURE

USING REAL GAS EQN. OF STATE:

Pwr'^(ecii<GUEss),T,)

SOLVE FOR COLD REGENERATOR DENSITY:

^REG

fP DENOM 1

1, = [-^ Fj^TOT -eci<GUEss)Vc,Jv„,-l

EVALUATE FUNCTION xp:

= "t" *6ci.(GUESS)' ~ "^y^REG 1, 2 /

" '^NEW YES

6cij(GUESS) ^ 6Clj(GUESS)

+

INCREMENTAL SEARCH

FIG. 4a. FLOWCHART FOR FINDING SYSTEM PRESSURE

36
An initial guess is then made for the instantaneous density in the cold expansion space:

pcli (guess) = 0.2 Pw-j (guess) (10)

XTl

The factor of 0.2 is present to insure that the initial guess of pel- is below the actual

value. Once pel. (guess) is known we can solve for a new pressure from a real gas

equation of state.

The equation of state that we will use is presented by Mann (2). The equation contains

seventeen terms and is contained in the Appendix for continuity. It is accurate from

ttK to 300K and to pressures of 100 atm. This equation is explicit in pressure; that is, pressure is found as a function (j) of density and temperature:

p = (pjy (11)

We can now substitute our first guess for the density in the cold expansion space along

with the temperature in this space in Eq. (1 1):

Pw^ = <^ (pel (guess), Tl) (12)

Next Eq. (6) is solved fot pregl in terms of p:l and Pw^:

pregli =|^Pwi^Denom

_ ^. ^^^^ _ ^^^^^^^ y^^^ y^eg^-l ^^3^

We now define a function ^ that will be zero when the appropriate denstities are

found:

* = 4.[pcli.Tl] - .fr [pregl, 1/2 (Tl+T2j] (14)

Note that ^ is really a function of pcli only because P regl can be found from

knowledge of p cL. We can now define an increment by which to sequentially increase

p cl| until we have found an interval of values of pcl^ that contain a root to Eq. ilk):

Ap = 0.28 <9"^ss) (15) m

We now increment pel. successively by Ap until the function changes sign. This

defines an interval whicVi contains a root of ^ and we can quickly converge on this root by the bisection method. The bisection procedure is illustrated in the flowchart shown

in Fig. 'fb.

This process is continued until the pressures and densities are known for each time

step in a cycle of the machine. The integrals for the time averaged pressure in the

working space and crankcase in equation 1 can then be evaluated numerically:

1-^ n 1=1

S: Pw = (1-Fj) Mtot R Tcr / Vcr (16)

37
Si = ©AVE Vl = Vave

®

(1>

Ci =

Cci,1OUESS

C2 =

6ci| GUESS Vi =

'''new V2 = VOLD

SOLVE FOR SYSTEM PRESSURE USING REAL GAS EON. OF STATE:

SOLVE FOR COLD REGENERATOR DENSITY:

Preg DENOM

''j'^TOT - Save V Iv "CljJ *REG 1

EVALUATE FUNCTION xp:

VaVE = (fiAVE- V - <t'(eREG•^^^)

YES

62 = Save V2 = Vave

CORRECT DENSITIES HAVE BEEN FOUND

fici, = Save

Pw,='^<eAVE'Ti)

FIG. 4b. BISECTION SEARCH

38
where we have assumed that the crankcase volume is essentially constant. The term

on the left represents the working space average pressure (Pw) while the term on the

right is the crankcase average pressure (Per). If the time averaged pressure is not the same in the working space we must adjust the mass fraction. This is done by relating

the pressures in the (j+l)th iteration to the pressures in the jth iteration as follows:

-pcrlj = krlj^, (l^Uil] (18)

We can solve for Fj+1 by substituting equations (17) and (18) into (16) , resulting in:

Fi+i - "Pw (19) ^ Fjpcr

After adjusting the mass fraction the progam goes through another cycfe and

calculates the system pressures for each time step. This prbcess is repeated until the

time averaged pressures in the working space and the crankcase are within some

specified error c.

The gross cooling for each stage is found by integrating the system pressure with

respect to the expansion space volume:

Qgl = a ^ pw^. Xi (20)

Qg2 = A2 Ogl

Al (21)

Qg3 = A3 Qgl

AT

(22)

These expressions are valid because the expansion spaces are assumed to be iso- thermal.; the cyclic integral of the work done by the gas is then equal to the heat

absorbed by the gas. "However, enthalpy ls a function of pressure and since

the ingoing and outgoing streams are at different pressures, this leads to

an incomplete account of the Joule-Thomson effect.

After finding the gross cooling for each stage the net cooling is found by subtracting

off the various thermal losses presented in Table 1:

Onet = Qg - Q loss (23)

The thermodynamic work done by the compressor is found by numerically integrating

the system pressure with respect to the compression space volume:

Ur = Az n • TT .11 PWi Zi (24)

The net work done by the compressor is found by adding the thermodynamic work to the flow by adding the thermodynamic work to the flow losses of the gas in the

transfer line and regenerators:

Wc(net) = Wc + Flow losses (25)

39
TABLE I - Thermal Losses

LOSS

Shuttle

CAUSE

Regenerator motion causes heat to be transferred from warm to cold

end.

Regenerator Irreversible heat transfer between regenerator matrix and helium.

Heat Transfer

Temperature Temperature cycling of regenerator matrix

Swing

Pressure Drop Viscous drag in flow passages causing reduction in gross cooling.

Conduction Irreversible heat transfer between warm end of displacer assembly and

cold end.

TABLE II - Summary of Optimized Three-Stage Cooler Design

a. Data for Various Stages

Stage

Temperature, °KSwept Volume, in- Dead Volume, in Regenerator Void Volume,

in Re^nerator Flow Area,

in Regenerator Matrix Material

Regenerator Length, in.

1

10

2.0'f X 10 8.16 X 10 3.70 X

10'

-3

-5

i^JH X 10"^

Lead-Antimony

Spheres

0.780

25

X 10

1.75 X 10

I. 16 X

10'

1.32 X 10"^

Nickle Spheres

0.880

123 - 7.73 X 10,

3.09 X 10"^

3.80 X

10""^

X 10"^

Nickle Spheres

0.930

b. Ambient End Data

Charge Pressure - 500 Psia

Ambient Temperature - 300K

Cycle Speed - 25 Hz - Compressor Swept Volume - 0.131 in _ ^ Compressor Dead Volume - 2.98 x 10" in ^ Connecting Line Dead Volume - 5.65 x

10" in

40
RESULTS

Airborne SQUIDS require small, efficient cryocoolers with small magnetic signatures. To

achieve the small packaging requirement designs utilizing high charge pressures and cycle

speeds to compensate for the lack of physical size were examined.

The optimized three stage cooler design is summarized in Table 2. The design was

optimized by adjusting the length of the regenerators and the stage temperatures until the desired net cooling was obtained at the coldest stage. The performance of the

machine is shown in Table 3.

The program was re-run using the same data except that the real gas equation of state was replaced with the ideal -gas equation of state. The results are shown in Table 4. It

is clear that the effect of compressibility in the cold stage expansion 'space and

regenerator substantially altered the performance of the machine. The pressure ratio

and gross cooling were reduced when compressibility was not included in the analysis.

Compresssibility effects tend to magnify the swept volume in the cold expansion space

relative to the other subvolumes in the machine. The compressibility factor 2 is defined by:

PV = ZMRT

The effective temperature weighted swept volume of the cold expansion space is then

given by: A Vcl = AVcl

Zcl Tl (27)

where Zcl is the cyclic average compressibility factor. In the cold expansion space,

similarly, the effective temperature weighted dead volume of the cold end regenerator

is: ^ Vregl = Vregl (28)

Tregl Tl

As can be seen from Fig. 5, the compressibility factors for both the cold end expansion

space and the cold end regenerator are always less than unity. However, the average

compressibility factor for the cold end expansion space is less than that of the cold

end regenerator (since the regenerator is at a higher temperature). This causes the

total temperature weighted swept volume of the machine to increase at the expense of the total temperature weighted dead volume. The pressure ratio and gross cooling are

thus increased.

Several runs were also made using various displacer motions other than sinusoidal.

Motions were found that gave substantial increases in net cooling. The compressor

swept volume was then reduced until the net cooling was close to its original value,

but because of the smaller compression space the input power was reduced. The

performance characteristics of this design are listed in Table 3 and the optimized

displacer waveform is shown in Fig. 6. The optimum waveform has short dwells at the ends

of the stroke that tend to round off the work curve in the cold expansion space. By

using the optimized waveform, the COP of the cooler was increased by 9%.

41
TABLE III - Performance Characteristics of Three Stage Cryocooler

Stage 1

Gross Cooling, W 0.972

Shuttle Loss, W 8.26 x 10" Regenerator Heat Transfer 0.130

Loss, W

Regenerator Temperature 0.804

Swing, K

Temperature Swing Loss, K 0.268 Pressure Drop Loss, W 0.283

Total Conduction Loss, W 'f.50 x 10 Pumping Loss, W 1.70 x 10 Net Cooling, W 0.220

-2

2

2.09

0.191

0.71'f

0.180

f .73 X 0.539

5.86 X

7.66 X

lAtt X

10'

-3

-3

3.68 0.320

O.'fOO

2.56 X

10'

2.35 X

0.488

2.43

3.51 X 6.55 X

10-3

Pmax - 46.1 atm

Pmin - 26.9 atm

Pressure Ratio - 1.71 Motor Shaft Power - 54.4 W

Flow Losses - 3.10 W

TABLE IV - Performance Characteristics of Three-Stage

Cryocooler Ignorring Compressibility

Stage 1

Gross Cooling, W 0.833

Shuttle Loss, W 8.26 x

10' Regenerator Heat Transfer 0.148

Regenerator Temperature 0.938

Swing, K

Temperature Swing Loss, K 0.337 Pressure Drop Loss, W 0.303

Total Conduction Loss, W 4.50 x

10" Pumping Loss, W 1.41 x

10' Net Cooling, W -2.30 x 10

2

2-2

1.79 0.191

0.646

0.190

4.99 X 10

0.570

5.86 X 10 8.34 X 10 -0.259

-1-2

3.16 0.320

0.366

2.63 X

10'

2.37 X 0.507

2.43 0.906

-0.497

10"

Pmax - 44.0 atm

Pmin - 27.5 atm

Pressure Ratio - 1.62 Motor Shaft Power - 52.2W

Flow Losses - 3.77 W

42
TABLE V - Performance Characteristics of the Three-Stage Cryocooler

with Non-Sinusoidal Dispiacer Motion

Stage 1

Gross Cooling, W 0.889

Shuttle Loss, W 8.27 x

10" Regenerator Heat Transfer 0.12^^ Loss, W

Regenerator Temperature 0.830

Swing, K

Temperature Swing Loss, K 0.271 Pressure Drop Loss, W 0.206

Total Conduction Loss, W t^.5Q x

10" Pumping Loss, W 2.11 x

10' Net Cooling, W 0.213

1.91 0.192

0.613

0.183

itAS, X 0.382

5.86 X

9.61 X

8.2^^ X

10"

10

10

10

-13

2

3.37 0.320

0.312

2.6'f X

10"

2.18 X 0.336

2A3

kAl X -7A7 X

10'

10"^

10"

Pmax - 'f8.0 atm

Pmin - 25.8 atm

Pressure Ratio - 1.86 Motor Shaft Power - 'f8.6 W

Flow Losses - 1.66 W

43
44
45
CONCLUSION

Tnis investigation demonstrates the feasibility of constructing a small, efficient three

stage lOK Stirling Cryocooler. Compressibility should be taken into account when

writing computer simulations for cryocoolers operating in this temperature range.

Also, °he displacer motion can be optimized to increase the performance of the

machine.

NOMENCLATURE

Az - Swept area in compressor piston (m ) ^

Al,2,3 - Swept area of expansion spaces 1, 2, and 3 (m ) Fj - Mass fraction in working volume for jth mass fraction iteration

Mcrj - Mass in crankcase volume in jth mass fraction iteration (kg)

Mtot - Total mass of helium charge (kg)

Mwj - Mass in working volume in jth mass fraction iteration (kg) n - Number of discrete time steps per one cooler period

P - Pressure (Pa)

Pc - Charge Pressure (Pa) Per - Crankcase Pressure (Pa) Pw. - Instantaneous working volume pressure (Pa)

Qgl,2,3 - Gross cooling in stages 1, 2 and 3 (W)

Qnet - Net cooling in a particular stage (W)

Qloss - Cooling losses in a particular stage (W)

R - Ideal gas constant for helium (J/kg K)

T - Temperature (K)

Tamb - Ambient temperature (K) Tcr - Crankcase temperature (K)

Tl, T2, T3 - Temperature in expansion spaces 1, 2, and 3 (K)

y - Volume (m^)

Vamb. - Instantaneous ambient end volume (m ) ^ Vcl, ^c2, 7c3 - Midstroke \u>lume in expansion spaces I, 2, and 3 (m ) Vcr - Crankcase volume (m ) - Vreg 1, 2, 3 - Void volume in regenerators-1, 2, and 3 (m ) Vzi - Instantaneous compressor volume (m ) Wc - Thermodynamic work done by compressor (W)

Xi - Instantaneous displacer velocity (m/s)

Zi - Instantaneous compressor velocity (m/s)

Z - Compressibility factor

Vcl; ,Vc2j,Vc3- - Instantcineous volume in expansion spaces 1,2, and 3 GREEK LETTERS

p - Density (kg/m ) , preglj - Instantaneous gas density in regenerator 1 (kg/m )-

pel. - Instantaneous gas density in expansion space 1 (kg/m ) T - Cooler period (sec)

SUBSCRIPTS

Associates a quantity with the ith time step

Associates a quantity with the jth mass fraction iteration

46
APPENDIX; Equation of State for He

h Mann (2) presents the following equation of state for He between 't.OK and 300K. This

equation shows good correlation with experimental data for pressures up to 100 atm:

P = RdT + (Rn^ T + n2 + n^/T + n^/T^ + n^/T'^)d^

+ (Rn^T + ny)d^ + n^ d'^

+ (n9/T2 + NjQ/T^ + njj/TV

+ exp (-nj2 d^/T)

+ (n^3/T^ + n^^/T^ + nj^^/T'*)d^

+ exp (-nj2 + n^^ d^ + n^^ d^

Where d is in gmol/liter, T in K and

R := 0.820575 (10"^) "9 = -0.5002871627(10"^)

"l

= 0.1627693557(10"^ "10 = 0.3'f51283837(10"b

"2 = -0.3700903<f92(10"^) "11 = 0.100892'tt38

"3 = 0.1567259965(10"^^ "12 = 0.65 (10"^)

\

= -0.1887566673(10"^ "13 = -0.1851562535(10"^

"5 = 0.'f7923't'f6't0(10"b

"l^f = 0.'f20'f351888(10"'^)

"6 = 0.15603't'f98't(10"^) "15 = -0.1712'ti^5150(10"^)

"7 = 0.927't2985'f2(10"^) "16 = 0.7739390202(10"^)

"8 = -0.28339970'f5(10"'*) "17 = -0.28706939^8(10"^)

BIBLIOGRAPHY

(1) S.B. Horn and M.E. Lumpkin, Advances in Cryogenic Engineering, Vol. I'f, Plenum Press,

New York (1974), p. 221.

(2) D. B. Mann, NBS TN 15^*, 1962.

47
SOME THERMODYNAMIC CONSIDERATIONS OF HELIUM TEMPERATURE CRYOCOOLERS*

D. E. Daney

Thermophyslcal Properties Division

National Bureau of Standards

Boulder, Colorado

Attempts to reach liquid helium temperatures with regenerative

cryocoolers have been generally unsuccessful. The principal and well

known cause of this failure Is the decay of the regenerator matrix heat

capacity at low temperature. As the regenerator effectiveness decreases, more and more of the refrigeration effect goes to the regenerator losses

rather than to the load. A less well -recognized problem for some cryocoolers has been operation under adverse thermodynamic conditions. For

example, operation at high pressure (40 bar) can decrease the calculated

efficiency of the lowest stage by a factor of 10 or more - primarily

because of the strongly negative Joule-Thomson effect of helium at low

temperature and high pressure.

We first Investigate the effect of the operating regime on the efficiency of helium temperature cryocoolers. Next, we propose two schemes

of hybrid regenerative-counter flow cryocoolers as solutions to the

problem of low helium- temperature regenerator heat capacity. Finally, we calculate the net mass flow through a typical hybrid regenerator-heat

exchanger for one of the schemes that we propose.

Key words: Cryocooler; hybrid cryocoolers; refrigeration; regenerator;

regenerative cooler; Stirling cooler; Stirling cycle.

1. Introduction

Much of the Interest In reliable, low power, and compact cryocoolers for cooling

smaller superconducting devices Is focused on regenerative devices such as the Stirling

machine. These coolers have a proven record of performance at temperatures above 20 K,

and their relative simplicity Is particularly appealing because 1t offers the potential

of high reliability. However, attempts to reach liquid helium temperatures with regenerative cryocoolers rejecting heat to ambient temperature have failed!. Although

temperatures as low as 6.5 K have been achieved, the efficiency falls rapidly at temperatures below 10 K [1,2,3]

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

1 Temperatures as low as 3.1 K have been achieved by Zimmerman and Sullivan [8] with heat

rejection at 8.6 K. This was achieved by operaton at low pressure - a very original

approach to alleviating both of the thermodynamic problems discussed In this paper.

48
The pHnc^ple and well known cause of th-fs failure Is the decay of the regenerator

matrix heat capacity at low temperature. An accompanying loss in regenerator effectiveness results in more and more of the refrigeration effect going into regenerator losses

rather than the load. A less well understood problem for some cryocoolers has been

operation under adverse thermodynamic conditions, i.e., operation at pressures and

pressure ratios where the efficiency is poor.

This paper proposes an alternative approach to the problem of diminishing low

temperature regenerator heat capacity. Instead of searching for new regenerator

materials, I suggest using hybrid cycles with helium counterflow in the final stage of

the refrigerator. Two schemes which embody this principle are given.

2. Regenerator heat capacity

The requirement of high regenerator heat capacity has long been recognized, and we

discuss it here only in sufficient detail to define the magnitude of the problem for

helium temperature regenerative coolers. An empirical expression for the effect of heat

capacity on regenerator effectiveness is given by Kays and London [4] for rotary regenerators: n 0.11 X ^0 JTY-^

where e is the regenerator effectiveness, eq the effectiveness for infinite regenerator

heat capacity and is the heat capacity of the regenerator divided by heat capacity of

the gas passing through in one direction during one cycle. Using this expression we find

that Cr must be greater than 3 for the degradation 'in e to be less than 0.01. By comparison, the experimental results of Gifford and Acharya [5] give greater than 5 for a lead

sphere regenerator operating between 80 K and 20 K. Whatever the exact values, both results support the conclusion that should be several times greater than one if a significant deterioration in effectiveness is not to occur.

Comparing the volumetric heat capacity of helium with that of lead (the best common

regenerator material), figure 1, we note that high losses may be expected below about 10 K. Although much may be accomplished through system optimization and materials research,

figure 1 offers little hope for efficient regenerative cryocoolers operating at 4 K. This,

of course, is not a new opinion.

3. Thermodynamic regimen

Because of their great complexity, regenerative cycle calculations are usually made

using the simplifying assumption of an ideal gas equation of state. So long as the operating temperature is well above the critical point, little error is generally introduced by

this assumption. As the operating temperature approaches the critical temperature, however,

the error introduced in some cases may be large enough to make the calculations meaningless.

In order to characterize the effect of the operating conditions on the performance of

the final stages we analyze a final Brayton cycle stage using real helium properties and

methods described previously [6]. Although not a substitute for detailed regenerative cycle

calculations, the results for a 6 K operating temperature, given in figure 2, clearly illustrate a strong effect of operating regimen on performance.

49
Figure 1. Volumetric heat capacity of helium and lead.

50
Figure 2. Performance of final refrigerator stage as a function of pressure and pressure

ratio.

51
The dependence of the ideal work on the pressure is pn'mar"fly due to the increasingly

negaffve Joule-Thomson effect of helium w^th increasing pressure. The result is progressively less cooling as the pressure increases. Low pressure ratios are less efficient

because a high fraction of the refrigeration is lost In the heat exchanger (regenerator).

For less effective heat exchange, the effect of pressure ratio Is even stronger. For the

model used here, at least, relatively low operating pressures and modest pressure ratios

are favored.

4. Hybrid regenerative coolers

As an alternative to the search for a high heat capacity regenerator material, we propose a hybrid machine with counterflow heat exchange In the final stage. Such an arrangement largely eliminates the need for a high heat capacity regenerator because heat Is transferred directly from one helium stream to another. Figure 3 shows two possible

arrangements for achieving such counterflow. In Scheme A, two regenerative coolers operating 180 degrees out of phase share a common counterflow heat exchanger, or hybrid heat

exchanger-regenerator In the final stage. The gas entering one expansion space Is cooled

In counterflow by the gas flowing from the second expansion space. The mass flow wave-form

Is distorted from the sinusoidal, so that some Instantaneous localflow Imbalance occurs.

Thus, a hybrid heat exchanger-regenerator (e.g. - a heat exchanger with high heat capacity)

Is desirable. Although a single pair of displacers are shown In the figure, any number

could be coupled In a similar fashion. For example, for three displacers the phase angle

would be 120 degrees and for four displacers the phase angle would be 90 degrees In order to

further reduce any flow Imbalance. Scheme A was suggested to me by Ray Radebaugh of NBS;

Its exact origin Is unknown to us.

In Scheme B the last stage of the displacer Is tightly sealed from the rest of the

system so that It may act as a conventional expansion engine. During the high pressure

part of the cycle, gas Is admitted to the plenum via valve 2. When the piston-displacer

reaches bottom dead center, valve 3 opens and admits gas to the cylinder until the piston

rises to the cut-off position. As the piston continues upward, the gas cools by expansion;

and a portion Is liquefied. At top dead center the exhaust valve (4) opens, and two-phase

helium Is discharged Into the load heat exchanger-plenum. Return of the gas to the main

working volume occurs through valve 1 during the low pressure part of the cycle. Counterflow heat exchange between the high and low pressure streams occurs In a somewhat conventional heat-exchanger. However, because the mass flow In the two heat exchanger streams Is not Instantaneously equal, one side (the low pressure side) should have a volume much

greater than the other side to provide additional heat capacity. Because of Its complexity.

Scheme B Is probably feasible only for coolers with a capacity near one watt and above.

Scheme B Is an extension of a concept suggested by S. C. Collins some years ago [7].

In order to calculate the reduction In the regenerator load accomplished by Scheme A, we calculated the pressure and mass flow for a 3-stage Stirling cooler using real helium

properties and sinusoidal movement of the displacer and piston. The following characteristics were assumed:

52
Regenerator

Final Stage

Hybrid Counter Flow

Heat Exchanger-Regenerator

Scheme A. Displacers 180° out of phase.

Valves

Plenum

Chamber

Counter Flow

Heat Exchanger

Regenerator

Piston Rings

Exhaust Valve

Inlet Valve

Scheme B.

Figure 3. Hybrid refrigerator schemes for the final stage of regenerative cooler.

53
Table 1. Split-Stirling cooler parameters for sample calculation of mass flow

Isothermal compressor temperature 300 K 1st expansion space temperature 100 K 2nd expansion space temperature 25 K 3rd expansion space temperature 6 K Compressor volume 300 cm3

1st expansion space volume 17 cm3

2nd expansion space volume 4 cm^

3rd expansion space volume 1 cm3

Regenerator void volume = 1/2 expansion space volume.

Figure 4 shows that the mass flow rate is nonsinusoidal with the flow maximum and mini- mum separated by the compressor-displacer phase angle, 90 degrees in this case.

It is the total cyclic mass flow through the regenerator, not the instantaneous flow

rate, that determines the load on the regenerator heat capacity. Thus, in Table 2 we com- pare the effect of the number of displacers on the cyclic change in the third stage expansion space mass.

Table 2. Cyclic mass flow with multiple displacers

Compressor-displacer angle

Number of 90 degrees 120 degrees

Displacers Flow ratio 1/N x flow ratio Flow ratio 1/N x flow ratio

1 1.0 1.0 1.0 1.0

2 0.66 0.33 0.38 0.19

3 0.25 0.12 0.03 0.01

4 0.06 0.02 0.01 0.004

The column denoted flow ratio is the cyclic change in the total third stage mass compared

to the change for a single displacer. Since multiple displacers would give N times the

capacity of a single displacer, a further reduction in mass flow is achieved by the reduction in the size of each displacer when multiple displacers are used. Thus, the column

denoted 1/N x flow ratio gives the total reduction cyclic mass flow for Scheme A.

A dramatic reduction in mass flow occurs when the number of displacers gives a phase

angle between displacers equal to the compressor-displacer phase angle. Four displacers

(with a 90 degree compressor-displacer phase angle) give a fifty fold decrease in the

regenerator heat load.

5 . Summary

In summary, I have proposed an efficient helium temperature regenerative cryocooler.

It is a hybrid device using counterflow heat exchange in the final stage. It should have

a maximum cycle pressure below 10 bar, and a pressure ratio above 2.

54
10

8

6

4

2

0

-2

-4

-6

I

IT

^

2ir

Figure 4. Pressure and mass flow waves for a Stirling cooler.

6. References

Stuart, R. W. and Cohen, B. M., Operation of a Three Stage Closed-Cycle Regenerative

Refrigerator in 6.5 K Region, in Advances in Cryogenic Engineering, Vol. 15, edited

by K. D. Timmerhaus (Plenum Press, New York, 1970).

Daniels, A. and du Pre", F. K., Triple-Expansion Stirling-Cycle Refrigerator, Advances

in Cryogenic Engineering, Vol. 16, edited by K. D. Timmerhaus, 178-184 (Plenum

Press, New York , 1971 ).

Zimmerman, J. E. and Radebaugh, R., Operation of a SQUID in a Very Low-Power Cryocooler, NBS Special Publication SP-508, edited by J. E. Zimmerman and T. M. Flynn,

National Bureau of Standards (1977).

55
[4] Kays, W. M. and London, R. L., -fn Compact Heat Exchangers, 20 (McGraw-H-fl 1 Book Co.,

Inc., New York, New York, 1978).

[5] G"ffford, W. E. and Acharya, A., Low-Temperature Regenerator Test Apparatus, In Advances in Cryogenic Engineering, Vol. 15, edited by K. D. Timmerhaus, 436-442

(Plenum Press, New York, New York, 1976).

[6] Daney, D. E., Low-Temperature Losses in Supercritical Helium Refrigerators, in

Advances in Cryogenic Engineering, Vol. 15, edited by K. D. Timmerhaus, 205-212

(Plenum Press, New York, New York, 1976).

[7] Daney, D. E., Design, Construction, and Testing of a Small Novel Air Liquefier,

S. M. Thesis, Massachusetts Institute of Technology, 1962.

[8] Zimmerman, J. E. and Sullivan, D. B., A milliwatt Stirling cryocooler for temperature below 4 K, Cryogenices 19_, 170-171 (1979).

56
DYNAMIC ANALYSIS OF A SMALL FREE-PISTON RESONANT CRYOREFRIGERATOR

R.A. Ackermann, Ph.D

Mechanical Technology Incorporated

Latham, New York 12110

The development of small, cryogenic refrigerators is receiving increased attention through the use of linear resonant machinery which

has the potential for increasing the reliability and efficiency of

these devices. The use of linear machinery, however, presents a complex

second dimension to the analysis of the system's dynamics and thermodynamics which are implicitly tied together through the pressure wave produced in the refrigerator. Tnis paper describes the dynamics of a linear cryogenic refrigerator and presents the analytical development

of the equations of motion through the use of phasor notation. Solutions for both the transient and steady-state motions are presented,

and the conditions of stability and selection of key operating parameters such as frequency, piston-displacer amplitudes and phase angles

are discussed.

Key words: Displacer and piston drives; linear resonant machinery;

linear free-displacer cryorefrigerator; linear electric motors; phasor

notation; thermodynamic and dynamic analyses.

1.0 Description of a linear resonant cryogenic refrigerator

A schematic for a free-piston linear resonant refrigerator is shown in Figure 1. The

machine consists of two dynamic components: a compressor piston which is electromagnetically

actuated at resonance by a linear electric motor, and a displacer which is actuated at reson- ance by the pressure wave developed in the machine. Refrigeration is produced in the device

by the expansion and compression of the working fluid at different temperature levels and by

the controlled flow of the working fluid between the temperature levels due to volume changes

in the expansion and compression spaces. The operation of the refrigerator is defined by the

following set of equations.

1 .1 Volume equations

The equations defining the changes in the expansion and compression volumes for sinusoidal motion of the piston and displacer are:

V(expansion) = = j (^5)^ C + sinut' )

V(compression) = = i (V^) (1 + sin (wf - e))

c 6 c 0

57
where:

(V ) = Swept volume in the expansion space

^ e 0

(V ) = Swept volume in the compression space

c 0

(i) = Operating frequency (rad/sec)

e = Angle by which volume variations in the expansion space lead those

in the compression space.

V = t + t^

Fig. 1 MTI Single-Stage Linear Free- Fig. 2 Displacer and Volumetric

Displacer Cryorefrigerator Phasor Diagram

1.2 Pressure equation

The pressure equation is derived from the mass continuity equation for the system and

is given by

58
where

Mj = Total mass in the system

P = Pressure

T = Temperature

V = Volume

R = Gas constant

subscripts

ch = Charge pressure

e = Expansion space

c = Compression space

V = Void volume

To = Initial charge volume

a = Ambient

The conversion of the absolute volume expressions to phasor notation is made through the

transformations:

e Zen ' 0

^

where:

V = Absolute volume variations

Vg = The phasor expansion space volume variation due to the displacer motion

and

v^ on

J

= Compression space volume variation due to the displacer moti

V = Compression space variation due to the piston motion

c J p

The advantage of this phasor notation is that the amplitude and phase relationship of

each of the system operating parameters may now be represented on a phasor diagram that pro- vides physical insight into the mathematics. The phasor diagram for a cryogenic refrigerator is shown in Figure 2. The development of this diagram proceeds from the relationship

between the two position phasors.

Piston Motion

x_ = (x ) sinwt

P P 0 Displacer Motion

= (xj) sin (wt + y)

59
where:

(x^)^ and (Xp) = the amplitudes of the motion

Y

= the angle by which the displacer motion leads the piston

motion.

From these motions, the expansion and compression volume phasors may be located on the diagram (Figure 2) as follows:

• The expansion volume variation is produced by the displacer and is 180°

out of phase with the displacer motion.

vt = -A,(x,) sin(wt + y) = A, (x.) sin oit' e a do dao

= the displacer frontal area

• The compression volume variations are produced by both the displacer and

piston motions.

V = V J + V

c c,d CP

where v^^p is the variation produced by the piston motion, which is represented by a phasor that is 180° out of phase with the piston motion

r, = (x„) sintot c,P p p 0

Ap = the piston frontal area

and Vcd the variation produced by the displacer, which is in phase with

the displacer motion

^c,d = (^d - ^OD^^^d^o

^^'^ ^ ^)

ApQpj = the displacer rod area as shown in Figure 1. The

function of the displacer rod will be covered in a later discussion.

The third dimension in the analysis is the development of the pressure phasor, which is the resultant of the pressure variations caused by the displacer and piston motions. This

relationship is shown in Figure 3, and is defined as

^

~ e,d "^cd '^cp

60
• The displacer motion produces two pressure effects. First, variations in the

expansion space produce a pressure change because of the temperature difference

between the expansion and compression volumes and the shuttling of the working

fluid between these volumes by the displacer motion. For a cryogenic refrigerator, where the expansion volume is at the cold temperature, the expansion volume component of the pressure wave is a phasor in phase with the displacer motion. This pressure component is represented by ?e,d- Secondly, the volume

variations produced by the displacer motion on either end of the displacer

differ by the volume occupied by the displacer rod. Thus, neglecting the temperature difference, the displacer motion alsa causes a pressure change 180°

out of phase with its motion. This pressure component is represented by ^.

• The piston motion produces a pressure change due to volume changes of the com- pression space. This motion produce^^ a pressure change in phase with the piston motion, which is represented by p. _ in Figure 3. c , p

The sum of these three components produces the ideal pressure phasor. In an actual maj chine a pressure drop will occur across the regenerator, producing a phase lag between presI sure changes in the expansion and compression spaces. The pressure drop is represented by

Ap in Figure 3, and the expansion and compression space pressure phasors are given by

P = (P ) sin (cot' + a)

e e o

P^ = (P^)^ sin (wt' - (3 + 6))

where

a = angle by which the pressure phasor (pg) leads the expansion volume phasor

B = angle by which the pressure phasor (P^) lags the compression volume phasor.

Fig. 3 Pressure Phasor Diagram Fig. 4 Dynamic Free-Body Diagram

61
Up to this point in the analytical development, the equations apply for both a kinematic or free-piston machine. In a kinematic machine the proper amplitude and phase separation of the piston and displacer are established by a system of mechanical linkages. In a resonant machine the mechanical linkages are replaced by an equivalent set of dynamic components (springs, masses, dampers, and a displacer drive). Dynamically, the free-piston, linear

resonant cryogenic refrigerator may be represented by the spring, mass, damper system shown

in Figure 4. The link between the piston and displacer is a pneumatic linkage created by the

engine pressure wave; thus, the additional complexity in analyzing this device becomes evident. Analysis of a kinematic device is less complicated because the linkage (drive mechanism) is an independent variable controllable through the positioning of the linkage on a crank mechanism. In a free-piston machine the pneumatic linkage is a dependent variable es- tablished by the dynamics of the system. Thus, the thermodynamics and dynamics are tied to- gether through the pressure phasor.

1.3 Thermal equations

Having established the pressure and volumetric phasor relationships, definition of the

power input, refrigeration capacity, and equations of motion for a resonant machine now are possible. An overall energy balance of the system gives the power delivered to the machine

as the difference between the heat rejected to ambient and the refrigeration capacity.

''in " ^REJ " ^REF " ^EM

'

where

^EM

~ electromagnetic driving force

Xp = piston velocity

Considering isothermal compression and expansion processes

%ir - ""e

By applying the following phasor relationships to these integrals,

dV = dv

P = + ^

where

P^ = mean system operating pressure

and integrating over a full cycle leads directly to the following expressions for the refrigeration and heat rejection

QrEF = -f-^d-(^d^o ^Pe^o

'^'^^

D 62
where

f = the operating frequency in cycles/sec

These two equations define the difficulty in designing a resonant cryorefrigerator.

First, the only independent variables are the frequency (f), and the cross-sectional areas,

Aj, ArqD' 3nd Ap. Second, the five variables (X(j)q, (xp)o, a, 6, and y are all dependent

variables that are functions of the dynamic components, i.e., the masses, spring components,

and damping. Finally, the pressure is a function of the temperatures and volumetric changes

as defined by the mass continuity equation. Equation 1. Therefore, the performance of a linear resonant cryorefrigerator is defined by a minimum of 3 simultaneous equations. These

are

1. The thermodynamic equations defining the refrigeration capacity

2. The mass continuity equation

3. The equations of motion

1.4 Equations of motion

The equations of motion are derived from the dynamic free-body diagram presented in

Figure 4. The equation of motion for the displacer is given by

Md^d * ^d^d "^d^d I'd = °

in which MdXd is the inertia force on the displacer, CdXd is the damping force which results

from the viscous losses in the gas springs, K^xj is the conservative spring force, and

Fd is the displacer driving force that maintains the displacer motion. The equation defining the driving force is found from a force balance on the displacer as shown in Figure 5.

This balance is

•"d " ""e'^d ~

''c^'^d " \0D^ " VrOD

In terms of the phasors described in Figure 3, this reduces to

where

AP = P, - Pg

p = P - P ^C C 0

Referring again to the phasor diagram in Figure 3, it may be seen that the pressure

phasors pg, p^, and Ap can be written as linear combinations of the conservative (displacement) and nonconservative (velocity) components as shown in Figure 6. The figure shows that

the following relationships exist.

63
Ad

Pe

OlSPLACER

Tcold

HEAT EXCHANGE COMPONENTS

warm

-Wl

kpXp

Ap/ CpXp L

\

Xp

Fig. 6 Pressure Phaser Components

Fig. 5 Displacer Force Diagram

Ap = -c^Xj + Cpip

Substitution of these expressions into the displacer equation of motion leads directly to ¥d ^ (^d ^ ^d-^d^^d - S'^^d'^p (^d ^ '^d-'^ROD^^d ^ W^p = °

in which the term cj-Ad-Xci represents a nonconservative restoring force on the displacer pro- duced by the viscous drag in the heat exchanger components; and Cp-Aj-Xp and kp-AROD'Xp

represent the driving forces on the displacer produced by the pressure drop across the displacer, i.e., the heat exchanger Ap, and the pressure force on the displacer rod.

The equation of motion for the piston is given by

M„x« + C„x„ + K x„ + F„ - Ft,„ = 0 PP PP PP P EM

64
in which the inertia, damping, and spring forces are described as above. The pressure force

on the piston is

p "^c p p p d d

and the electromagnetic driving force for the motor is given by

^EM = ^Im(*)

where:

a = e/ip

e = Voltage drop across the motor

I (t) = Current in the motor circuit shown in Figure 4.

For our case, where the electromagnetic driving force is produced by a linear induction motor, the current is found from

or, by differentiating

L

^ + R-j^ + 3 " EQwcos(a)t + n)

dt

where:

L = the motor inductance

R = internal electrical resistance

Eq = the amplitude of the applied voltage E(t)

0) = the operating frequency

n. = the phase angle between the applied voltage and the piston displacement

Then, the three equations defining the system dynamics are

"^d 0 0 ^d

0 ^d

0 ^d

0

0 "p 0 + 0 S

0 + ^

a ^P

0

0 a

L 0 0 R 0 0 0 FQa)COs(a)t + n)

where:

^d = ^d " ^d-'^d

"^d = ^d ^ ^d-'^ROD

65
The solution to the dynamic equation, in matrix form, is

X = +

in which x^ represents the transient solution, and Xg represents the steady-state solution.

A discussion of these two solutions follows.

1.4.1 Transient solution

The transient equation is

^d

0 0 ^d ^d

0 ^d "^d 0 ^d

0 M

P 0 XP + 0 S

0 y

+ ^d

R

P

a X

P 0 a L ^m 0 0 R m

0 0 0 I m

In the matrix form given, the transient equation represents a typical eigenvalue problem in which the eigenvalues represent the principal modes of vibration, and the eigenvectors define the amplitudes of the frequency response curves. In solving this set of equations the

electromagnetic force coefficient, a, in the mass and spring matrices prevents a closed-form

solution from being found and numerical techniques must be used to find the eigenvalues. In order to simplify the problem and gain some insight into the dynamics of the mechanical system, we shall assume that the electromagnetic force may be represented as a forcing function

of the form

"^EM "

""o ^'"(^^ ^)

where

n

= the phase angle between the force and piston motion.

With this assumption the transient equation becomes

'^d °'

+

"^d +

^d

0 Mp X

_

P_ 0 %

X_ p_^ s y _^p.

= [6]

Also, by drawing the force diagram for fhe displacer (Figure 7, which is the phasor representation of Equation 2), the term CpAjXp can be expressed as

66
This substitution leads to a closed-form solution and the following eigenvalues

- yosy) /(Kd - CpSiny - "kp) /^d

- V°Sy\^

"3,4 zMp^^y Mp V2Mpy

From necessary criteria for dynamic stability, the system is stable if the eigenval

coefficients

a. = -a. + i TbT

J J J I

satisfy the following conditions

If aj is negative the solution is unstable

If aj is positive the solution is stable

If bj is positive the solution is harmonic

-u/t

|Cp-Ad)

//

Xd _„^/

/(Kd + kd.ARODl Xd

Md'xV / a / \

(k(r*ROD) Xp

802640

Fig. 7 Displacer Force Phasor Diagram

67
If bj is negative the solution is not harmonic, and

If bj is negative and aj<bj the solution is unstable.

Thus, we see that in order for the mechanical system to be harmonic and stable, the neces- sary conditions are

which essentially says that the displacer driving forces due to the pressure differential across the heat exchange components and the displacer rod must not exceed either the conser- vative spring or nonconservative damping forces.

1.4.2 Steady-state solution

The steady-state solution is found by substituting for the motions

Xp = (Xp)^ sincot

x^j = (xj)^ sin(a)t + y)

This substitution leads to the steady-state equation

0 M_

-0) (x.) sin(a)t + y)

<J 0

-0) (x„) sinwt

^ P 0

d p +

)(x.) COs(a)t + y)

a)(x„) COSut

P 0

(x.) sin(a)t + y)

<J 0 0 V

(x„) sinwt

L P 0 FpSinCwt + n)

The solutions to this system of equations are given in terms of the piston-displacer ratio

and the phase relationship between the piston and displacer. The solution for the amplitude

ratio is

4 +

where:

2 _\

'dd M. the natural frequency of the displacer

68
= damping coefficient

And the solution for the phase angle is

Tany

2.0 Conclusions

The phasor analysis described presents a useful tool for analyzing and understanding

the operation of small, free-piston Stirling refrigerators. The usefulness of the approach

has been demonstrated by the insight it has given to the operation and selection of important

dynamic components. However, several assumptions were made in the analysis that have a bearing on its accuracy. These were:

1. The motions were described by a set of linear differential equations in which

the conservative force terms are linear functions of displacement and noncon- servative terms are linear functions of the velocity. Nonlinear effects are assumed negligible.

3. Only the first harmonic of the motions was used to describe the piston and

displacer motions. Higher order terms were considered negligible.

3. To simplify the analysis, it was assumed that the electrical' circuit could

be represented by a sinusoidal forcing function on the piston. The effect of

this assumption would be to decouple the motor's electrical circuit from the

machine dynamics, making the mathematics manageable. The operation of a real

machine is effected by this coupling where the interaction will influence

the dynamics of the piston and displacer.

The effect of nonl inearities on the equations of motion is the possible introduction of additional instabilities into the dynamic behavior; the effect of higher order terms is to introduce additional terms into the steady-state solution that may be of the same order of

magnitude as those given. A complete and thorough analysis of the device, therefore, must

rely on numerical procedures, and these procedures must integrate the dynamics and thermodynamics of the refrigerator.

[1] Walker, G., Stirling-Cycle Machines , University of Calgary, Canada, 1978.

[2] DeJonge, A.K., "A Small Free-Piston Stirling Refrigerator," 14th Intersociety Energy

Conversion Engineering Conference, American Chemical Society, Washington, D.C., 1979.

References

69
METHODS FOR THE MEASUREMENT OF REGENERATOR INEFFECTIVENESS*

Ray Radebaugh, Del Linenberger, and R. 0. Voth

Thermophysical Properties Division

National Bureau of Standards

Boulder, Colorado 80303

The regenerator ineffectiveness is usually the dominant heat loss

term in regenerative-cycle refrigerators which operate below about 15 K. Systematic measurements of regenerator ineffectiveness have never been

done in this temperature range in spite of the fact that theoretical

models are usually not adequate to fully describe and predict the thermal behavior of regenerators. The paper introduces the proper definition of regenerator ineffectiveness, which is valid for non-ideal gases

and systems where a pressure wave exists. All parameters which

influence the regenerator ineffectiveness are discussed. The concept of

a temperature- dependent ineffectveness is proposed which simplifies

computations of regenerator ineffectiveness when the material properties

are temperature dependent and which yields the average ineffectiveness

when it is integrated between the lower and upper operating temperatures

of the regenerator. Previous methods used and proposed for the measure- ment of regenerator ineffectiveness at cryogenic temperatures will be

reviewed and their limitations discussed. A new method will be presented which for the first time will allow measurements of regenerator

ineffectiveness at all phase angles between the mass flow rate and the

pressure wave, an important consideration with non-zero regenerator volumes. Temperatures considered are those beween 4 and 35 K.

Key words: Cryogenics; Gifford-McMahon cryocoolers; heat capacity; heat

transfer; helium; refrigerators; regenerators; Stirling cryocoolers;

thermal conductance; Vuilleumier cryocoolers.

1. Introduction

Small cryocoolers for temperatures down to about 7 K generally use a regenerative

cycle. The most common examples are the Stirling, Gifford-McMahon, and Vuilleumier cryocoolers. Their potential for long life at relatively low capacities is one of the main

attractions of the regenerative-cycle cryocoolers. They also have high overall efficiencies

for temperatures down to about 15 K. At lower temperatures the efficiency drops rapidly

because of the low specific heat of the regenerator matrix at those low temperatures.

Presently, the lowest no-load temperature which can be achieved with these regenerative-cycle

refrigerators is about 6 K, although Zimmerman and Sullivan [1] have operated a low-pressure single-stage device at 3.1 K with an upper temperature of 8.6 K. Higher pressures are required when the upper end is at room temperature, and the resulting higher heat capacity

of helium gas in the system can become comparable to or greater than the matrix capacity for

temperatures below about 10 K. Thus the overall performance of these refrigerators around

10 K and below is determined primarily by the effectiveness of the regenerator in transferring heat from the incoming warm gas to the outgoing cold gas.

* Work sponsored by Flight Dynamics Laboratory, Wright-Patterson Air Force Base, Ohio.

70
The theoretical behavior of regenerators is so complex that various simplifying assumptions are made in order to determine their behavior. Many of these assumptions are unrealistic for regenerators operating below about 15 K. Reliable experimental measurements of

regenerator effectiveness have not been done at temperatures below 20 K. Generally comparisons between different regenerators are done by substituting them in an actual refrigerator

and comparing the refrigerator performance. Such an approach is not very satisfactory

because (i) several variables may be changed simultaneously, (ii) the range of variables is

often limited, and (iii) it does not permit a separation of the various components which

contribute to the regenerator ineffectiveness.

In this paper we describe a method for measuring regenerator ineffectiveness which

allows the regenerator to be subjected to realistic operating conditions. These conditions

can be varied independently in order to better understand the behavior of a regenerator. In addition, the method allows for the separation of various components of regenerator ineffectiveness.

2. Definition and Theory of Regenerator Ineffectiveness

The effectiveness e of a heat exchanger is usually defined as

^ _ _ actual heat transfer rate

~

X

~ maximum possible heat transfer rate*

^max

The heat exchanger ineffectiveness x is given by

A = 1 - e = ^^^^ = Jll, (2)

Q 6 max ^max

where Qf-eg defined as (Qmax - Q)- The term Qy.gg then represents the heat load on the

region of interest (the cold expansion space) caused by the excess enthalpy flow into

this region. For a heat exchanger with a given A we then find Q^gq from

6 = A 6 . (3) ^req ^max

The introduction of the terms e and A are generally useful since they are either independent

of the temperature difference from one end of the heat exchanger to the other or a weak

function of it. Thus the terms e and A are useful in the characterization of a heat

exchanger since they eliminate one of the variables. The effectiveness according to Eq. (1) can be written as

^ ^ ^S^^.in - "^h,out^

{mC)„. (T. . - T^ . )'

mm h,in c,in

where C is the fluid specific heat at constant pressure and m is the mass flow rate. The

subscripts h and c refer to the hot and cold streams, respectively, 'in' and 'out' refer

to the two ends of the heat exchanger, and (mO^ip refers to the minimum of liicCc and my^C^

When C^and Cc are temperature dependent, averaged values must be used, or the equation

e = " " "jl" ^ h,out

{m[H(T, ,_) - H(T. . )!} .

can be used where H is the fluid enthalpy.

71
For many regenerative heat exchangers the mass flow rate, temperature, and pressure are

all functions of time. In that case Eqs. (4) and (5) can still be used for the regenerator

effectiveness but now e will be a function of time. In addition we must specify where

and ific are to be measured since they will vary with position in the regenerator for the same instant of time, unless the qas volume in the regenerator is negligible compared with the

gas volume passing through it. The most logical approach would be to choose m at the end of

the regenerator next to the region of interest, i.e. the expansion space.

Usually the instantaneous effectiveness is of little interest. Instead a value

averaged over a complete cycle is used to determine the average Qreg* The time averaged

effectiveness is then

^. ^^W\,J - V^h,out^^^^

^

where v is the cycle frequency. In this expression m^ = 0 during that part of the cycle

when cold gas flows from the expansion space through the regenerator. Likewise i%in would

be zero in part of the cycle. An energy balance on the regenerator requires that the heat

transferred from the hot stream equals the heat transferred to the cold stream. Hence, the

numerator in Eq. (6) could be replaced by the cold stream values.

We can simplify Eq. (6) whenever isothermal conditions exist for the gas inlets. In that case we let -jp = Ty and Jq -jp = Tl, where Ty and Tl are independent of time.

Equation (6) then becomes

^ _ vjrm,[H,(T^) - H,(T,^^^^)]dt

V J {m[H(T^) - H(T^_)]}^.^dt

and the ineffectiveness becomes _ --^W^,out^ -H,(T,_)]dt

vJ?r{m[H(T^j) -H(Tl)]}^,.,

when the denominator refers to the hot stream. Additional terms arise in the numerator when

the denominator refers to the cold stream and when the specific heat in the two streams are

not equal. The terms H(Tu) and H(Tl) are independent of time if (a) the pressure is con- stant, or (b) the gas is an ideal gas since H is independent of pressure for an ideal gas.

If either of those conditions is true, then Eq. (8) becomes

. . - -^W^h,out^ -H,(T^)]dt

V {m[H(T,) -H(Tl)]}^,.,

where m is the total mass of gas passing by the end of the regenerator. The mAH product

should be the minimum of ittcAHc and m^AHh, where

niu = J^iTiudt; m„ = m^dt. (10)

h h e c

Except for the case where some of the gas is being withdrawn (e.g. for producing liquid)

mh = m^.

For regenerators with a finite gas volume and which are subjected to pressure variations, Eqs. (7) and (8) are still not satisfactory. Consider the case where Tu = Tl in Eq. (8). The denominator then becomes zero and we would expect that the numerator would

72
also become zero in a manner such that x remains finite as Tu-»Tl. However as the gas pressure in the regenerator changes it will change temperature and therefore Th^out will

not be the same as Tl- Since the numerator in Eq. (8) remains finite as Ty-'TL, the

ineffectiveness approaches ± Thus in such a regenerator there is an additional source of

enthalpy flux besides the normal flux due to imperfect heat exchange in a temperature gradient. This additional enthalpy flux can be thought of as another heat flow term, but its

presence still indicates imperfect heat exchange within the regenerator. This additional

heat flow tenn can be understood by looking at the energy balance equations for the regenerator as shown below.

Gas Energy Balance Equation :

Heat Enthalpy Energy

transfer change storage

Heat transfer coefficient

Surface area per unit length

Matrix temperature

Gas temperature

Mass flow rate

Gas enthalpy

Cross- sectional area of gas stream

Gas density

Gas internal energy

Matrix Energy Balance Equation :

ha, {T„-T„) = A„ ^( p^U„)

L g m m dt m m

^ m m m 8t m

Heat Energy

transfer storage

An, = Cross- sectional area of matrix

p^, = Density of matrix

U„, = Matrix internal energy

= Matrix specific heat

We note that if the gas energy storage term were zero, i.e. Ag = 0, then a zero enthalpy

gradient and zero temperature difference between Tg and T^, satisfy both equations. How- ever when Ag ^ 0, the gas energy storage term will be finite whenever the pressure is a function of time. The finite gas energy storage term then forces all the other terms in

both equations to be finite even for Tg = Tl- This effect becomes particularly important

at low temperatures where the volumetric gas internal energy, pU, is significantly larger

than the matrix volumetric internal energy, PmU^,.

In order to properly utilize the previous definitions for e and a, we must separate

Qreg """^o *wo parts

^reg = ^T ' %' (^3)

where Qj is the heat load on the expansion space resulting from a temperature gradient in

73

h

= \

= Tm = Tg = m =
the regenerator and Qjj is the heat load caused by the qas energy storage effects in the

regenerator, which occurs when pressure waves exist. Figure 1 shows the expected dependence

of Q^eg 0" quantity {iTiC)min AT.^where AT is the temperature difference across the

regenerator. We would expect that % would depend on the ab sol ute. temperature and not on at for small at/T. Thus the intercept in Fig. 1 should give and Qj would then be Q^-eg - QuWe shall then consider the regenerator ineffectiveness to be that associated only with Qj.

The average ineffectiveness '•s just the slope of the straight line passing through the

point P in Fig. 1. Note that according to Fig. 1, changes somewhat with AT. This

behavior would occur when the thermal properties of the gas and the matrix are temperature

dependent. In addition, also depends on both Ty and Tl when Ty - Tl is large. Thus it

is difficult to characterize a regenerator in a general sense with the term Xg.

Figure 1. The dependence of Q^-eg on the parameter (mOmin^T. The intercept gives Qj. the

heat load on the expansion space due to qas energy storage effects in the regen- erator. The slope of the dashed line gives the average regenerator ineffectiveness at the point P.

We suggest that a regenerator is better characterized by using an incremental effectiveness e-j or ineffectiveness x^-, which is a function only of temperature. In Fig. 1 the slope at any point along the curve of Qj vs. (mC)niin is then x^-. For some fixed Tl,

e.g. 6 K, x^- is then evaluated as a function of absolute temperature and the average

ineffectiveness for a regenerator operating between any Tl and Ty is then

A.dT/(T^j - T^) (14)

whenever Cf, and are constant. When Cj, and Cc are not constant the X-axis in Fig. 1 should be replaced by m[H(Tu) - H(Tl)] and the average ineffectiveness becomes

74
^a=-^ S-C„,indT/[H(T^j) -Hd^)]. (15)

'l

A regenerator is then best characterized by specifying Qu(T) and A.j(T). Of course each

of these quantities will also be functions of the heat capacity ratio, dead volume, pressure

ratio, heat transfer units, pressure, and phase angle.

3. Measurement Methods

Figure 2 shows the schematics for three different methods of measuring regenerator

ineffectiveness. Figure 2(a) shows the dual regenerator technique developed by Gifford

[2,3]. In this scheme the gas first flows in one direction through the regenerators and

then reverses direction after a period of time due to the rotary valve. The pressure is

always constant in such a scheme. The heat flow due to ineffectiveness in the two identical

regenerators causes boil -off of some cryogen. Gifford used liquid H2 for the cryogen and so the value for Ti was fixed at 20 K. For lower temperatures liquid He can be used but

then a semi-weak thermal coupling [4] must be placed between the heat exchanger and the

helium bath so that the heat exchanger temperature will be high enough to prevent liquifying

the working helium gas. The method used by Gifford and in that proposed by Walsh et al . [4]

fixes the upper end of the regenerators at 77 K. The boil -off rate of the helium gives

the time averaged value of the heat flow, which is just Qreg» numerator of Eq. (8). A measurement of T^ out would yield the instantaneous x. This dual regenerator technique

is simple to carry out but its disadvantage is that the pressure is constant, unlike that

which exists in an actual refrigerator. Thus there is no way to study the effects due to gas energy storage.

Compressor

(a) (b) (c)

Figure 2. Schematics of three different methods of measuring regenerator ineffectiveness:

(a) the dual regenerator method, (b) the surge volume method, and (c) the variable expansion space method proposed here.

75
Figure 2(b) shows a scheme used by Moore [5] whereby a single regenerator is closed off

by a surge volume. That approach allows for a pressure wave within the system but the phase

angle between the mass flow rate and the pressure wave is fixed at -90°. In a real refrigerator the phase angle is close to 0°. The gas energy storage contribution to the effectiveness is phase angle dependent and so the approach shown in Fig. 2(b) still does not allow for a proper characterization of regenerators with finite gas volumes.

The method shown in Fig. 2(c) is proposed here as one which will provide a measure- ment of the regenerator ineffectiveness and under conditions which exist in an actual

refrigerator. The proper phase angle between the pressure wave and the mass flow rate is

established by a variable expansion space, which in this case is done with a displacer.

Heat supplied by the heater is absorbed by the refrigeration which occurs with proper phase

angles. The regenerator ineffectiveness causes a helium boil -off. Both heat exchangers

(H.E.) in Fig. 2(c) must be at the same temperature but yet they must be thermally isolated

from each other. This approach allows the refrigeration power to be separated from the

regenerator ineffectiveness. Figure 3 shows how this separation occurs. In this figure the

heat exchangers are now isothermal izers since they also serve to bring the gas passing

through them to the fixed temperature Tl. The cyclic nature of the process means that the

isothermal izers must have a high heat capacity. In order to simulate Stirling and

Vuilleumier cryocoolers the continuous compressor and rotary valve shown in Fig. 2(c) can be

replaced by a compressor which operates synchronously with the displacer.

Regenerator

Displacer

'reg 'ref

Figure 3. System boundaries for the energy balance equations which give Q^eg Qref"

An energy balance on the left isothermal izer shows that

^req

" - H(T)]dt, (16)

where m is positive for gas flowing to the right. For m positive T ^ Tl because of

the regenerator ineffectiveness. For m negative, T = Tl for a perfect isothermal izer

because the gas is leaving the isothermal izer . Thus mrH(TL) - H(T)] is zero when m is neg- ative, which means that Eo.^(16) is the same as the numerator in Eq. (9) if we allow for a sign change in Qpeq*. ^I^^s Q^gg is a direct measure of the heat flow due to regenerator

ineffectiveness and Q^gf is a measure of the refrigeration rate. The ineffectiveness is then given as

76
^reg . (17)

V MH(T,) -H(Tl)]}^,.,

An energy balance around the right isothermal izer and the displacer shows that

Q^gf

= vC^SVdV - jfrnHd^ldt], (18)

where P is the gas pressure at the displacer and dV is the volume change.

4. Description of Present Method

A detailed schematic of the low temperature parts of the proposed regenerator test

apparatus is shown in Fig. 4. Each isothermal izer is connected to a helium bath via a variable thermal link, which will be discussed in more detail later. The variable thermal links

allow the isothermal izers to be maintained at any temperature between about 5 K and 35 K with varying amounts of heat dissipation (up to 2 W) within each isothermal izer. It would

be possible to maintain isothermal izers #1 and #2 at 77 K to provide a comparison with

results from tests where the upper end is at 77 K. The heat dissipation from isothermal izer

To Compressor

Vonl

Iso- 'thsrmaNzor

1 *?

Regenerator

m

To Main

He Bath

To Mailt He Bath

® ® ®

O High SpMd SMWor

n Low SpMd SwMor

V HMtw

.j^Varlabl* TiMmMl Link

Displacer

Vent

a D Seal

9

o 5«

ISO- ^1 T) ffO (pjthermafceijr^Sgw

®

To Main

He Bath

isolated

He Bath

Figure 4. A schematic of the low temperature part of the regenerator test apparatus proposed here which shows the placement of m, P, and T transducers.

77
#3 is Qf-eq and so improved resolution is provided by using an isolated and shielded helium

bath to absorb the heat. A small amount of heat supplied by a heater on each isothermal izer

is used for temperature control. Isothermal izer #1 is included so that heat input to isothermal izer #2 gives O^eg on the hot side of the regenerator. The facility has the

capability to make (Tu - Tl)/Tl « 1 in order to measure incremental ineffectiveness at all temperatures between 5 and 35 K. In addition the temperature gradient can be reversed so that Qreg is measured by applying heat to isothermal izer #3. A heat input can be measured

much more accurately than a helium boil -off rate caused by heat dissipation. The change

in the sign of dT/dx through zero also allows us to determine %, the heat load due to gas energy storage.

High speed sensors within the gas stream at several places allow us to measure the

instantaneous T, P, and m for frequencies up to 20-30 Hz. Slow speed germanium thermometers

on the i sothermal izers are used to measure their temperature. Temperature control of the

isothermal izers is critical to the operation of the system. For measurements of incremental

ineffectiveness at, for example, 10 K the temperature difference Tu - Tl may be about 1-2 K. For an ineffectiveness of 2% the difference of T - T|_ may be about 20-40 mK. Thus temper- ature control to within ± 1 mK at 10 K is needed in order to measure the ineffectiveness to

within ± 5%. Temperature control of isothermal izers #2 and #3 is done in the usual fashion

using PID controllers and carbon thermometers as sensors. I sothermal izers #1 and #4 are to

be maintained at the same temperature to within ± 1 mK as their adjacent isothermal izer.

However, the calibration accuracy of germanium thermometers is not good enough to yield tem- perature differences less than 1 mK at 10 K. We thus propose to use a differential gas thermometer. Each i sothermal izer has a gas bulb of nominally the same size. The capillary

lines from each pair of gas bulbs go to opposite sides of a sensitive differential pressure

transducer. The room-temperature volume on one side of each transducer is made adjustable

so that the low- temperature volume to room- temperature volume ratio on each side of the

transducer can be made exactly equal. The correct volume is determined by a calibration at room temperature and at liquid He temperature. For a volume ratio of 0.05 and a filling

pressure of 1 atm at room temperature a 1 mK change in temperature causes a pressure differ- ence of about 2.7 Pa (0.02 torr). The room temperature volumes must be kept to within about

± 0.01°C of each other to keep differential pressure fluctuations less than the 2.7 Pa (0.02

torr). The electrical outputs from the two differential pressure transducers are fed into

separate PID temperature controllers, which in turn control the heat input to i sothermalizers #1 and #4.

The two vents shown in Fig. 4 allow the apparatus to measure the regenerator thermal

conductance, heat capacity, and heat transfer coefficient. To measure the thermal conductance, heat is applied to the right side of the regenerator and the heat is absorbed at the

left side by the gas stream flowing continuously through i sothermal izers #1 and #2 and out through the vent. The heat capacity of the regenerator (including that due to adsorbed He

and boundary layer He) is measured by passing He gas continuously through the regenerator

and out through the vent next to the displacer. After the gas and regenerator are in thermal equilibrium, heat is suddenly applied to the gas upstream from the regenerator to pro- vide a quasi-step function in temperature vs. time. The gas temperature at the regenerator

inlet, T^- , and the outlet, Tq, are then measured as a function of time. Figure 5 shows

the typical behavior for these two temperatures. The time lag is that associated with the wave front passing through the regenerator. The total heat capacity of the regenerator is given by

(19)

where Cp is the specific heat of the gas at constant pressure. The number of heat

transfer units, N^^, is defined as

^tu = ^'^/'^S'

(20)

where h is the heat transfer coefficient and a is the heat transfer surface area. The term N^u can be determined from the temperature- time curves in Fig. 5 from the relation

78
TIME

Figure 5. A temperature vs. time plot for the regenerator inlet and outlet temperatures

after a quasi-step change in temperature. The regenerator heat capacity and are derived from the temperature behavior.

^tu = ^^^i

-

High values of Ntu will be difficult to measure accurately since aTq will tend to be

smeared out in practice. If o is known, then h can be determined from Eq. (20).

The two critical parts of the regenerator test apparatus are the variable thermal links

and the isothermal izers. For the variable thermal links we propose to use gas gap heat

switches like that described by Bywaters and Griffin [6], in which heat is transferred

between several concentric copper cylinders by helium gas in the gaps. Figure 6 shows the

arrangement. For pressures below about 130 Pa (1 torr) the thermal conductance across the

helium gap is proportional to the gas pressure. The variable pressure is achieved by the

use of a molecular sieve adsorbant which can be heated to some temperature above 4.2 K [7].

When the molecular sieve is held at about 20 K by a heater, it adsorbs little He gas and so the pressure remains high. As the heat to the molecular sieve is decreased, the temperature

of the sieve decreases and the resultant adsorption causes the gas pressure to decrease. In the full 'off condition the only heat leak is through the fibergl ass-epoxy support tube.

Below 35 K radiation is negligible for polished surfaces. Measurements of the electrical

conductance between the two ends would show whether the copper cylinders are touching each

other. For long life (i.e., time greater than a few months) it is necessary to imbed a thin

foil of Cu-Ni or stainless steel within the fibergl ass-epoxy tube in order to prevent the

diffusion of He gas through the tube walls.

The isothermal izers are the other set of critical components. Their design is not easy

because of the following four requirements: a) high heat capacity, b) high N^^,

) c) small pressure drop, and d) small gas volume. The first two and the last two requirements generally work in opposition to each other. The temperature difference between the

gas and the isothermal izer at the outlet, aTq, is related the inlet temperature difference, AT-j , by

aTq/aT. = exp(-N^^). (22)

-4

Thus for aTq/aT-} = 10 , we need N-tu = 9.2. With such values the gas outlet temperature

is within 1 mK of the isothermal izer even for aTq = 10 K. To calculate the required heat

capacity of the isothermal izers we estimate the total heat to be stored in each cycle.

79
Figure 6. Drawing of a gas-gap heat switch used for the variable thermal link. The helium

gas pressure between the concentric copper cylinders is controlled by the temperature of the molecular sieve adsorbent.

For a refrigeration rate of Q^ef the total heat to be stored is Q^ef/v. For a total heat

capacity of C-j the temperature fluctuation 5t in the i so thermal izer is then given by

C,.«T = Q^g^/v. (23)

For a fixed Q^gf the largest C-| is needed at the lowest v. For Q^ef = 1 W and C^- = 50 J/K we get 6T = 5 mK at v = 4 Hz but 6T = 1 mK at v = 20 Hz. Further reductions in 6T can be

achieved by staging. A total heat capacity of at least 40 J/K between 5 K and 40 K can be

achieved with the use of 50 cm^ of Pb and 100 cm^ of He gas at a pressure of 2 MPa (20

atm). The dimensions of the Pb and the He space must not be larger than the thermal penetration depth at the working frequency range, which at 10 Hz is about 1 mm in Pb at 35 K and

50 ym in He below 20 K.

80
Design calculations show that an N^^ value of 9.2 can be achieved with pressure

drops generally less than 1% of the pressure for pressures greater than about 1 MPa (10 atm)

land temperatures up to 35 K. A flow rate of 1.5 g/s can be accomodated with a gas volume in

the isothermal izer of about 0.3 cm^. We propose to use an annular gap of about 28 ym

thickness, 31 cm effective width, and 3.5 cm long to provide the necessary surface area.

I 5. Conclusions

i The regenerator test apparatus proposed here should be capable of measuring regenerator

1 ineffectiveness in the range 4-35 K with variable phase angle between the mass flow rate and

j the pressure wave. Thus the conditions can be made to match very closely that which occurs

! in an actual refrigerator. In addition, the apparatus would be capable of separating out

all of the sources contributing to the overall ineffectiveness and show their dependence on various parameters.

6. References

I [1] Zimmerman, J. E. and Sullivan, D. B., A Milliwatt Stirling Cryocooler for Temperatures

I Below 4 K, Cryogenics 19, 170-171 (1979).

I [2] Gifford, W. E., Basic Investigation of Cryogenic Refrigeration Methods, Air Force

Flight Dynamics Laboratory, Wright-Patterson Air Force Base, Ohio, 45433,

AFFDL-TR-68-61 , (1968).

[3] Gifford, W. E. and Acharya, A., Low-Temperature Regenerator Test Apparatus, Adv. Cry.

Eng. 15^, 436-442 (1970).

[4] Walsh, P. J., Wade, L. A., and Chaudhary, T. M., Cryoregenerator Heat Capacity

Enhancement, Air Force Flight Dynamics Laboratory, Wright-Patterson Air Force Base,

Ohio, 45433, AFFDL-TR-80-3034 (1980).

[5] Moore, R. W., Jr., An Investigation of Thermal Regenerators for the 4 to 20 K Temperature Range, Air Force Flight Dynamics Laboratory, Wright-Paterson Air Force

Base, Ohio, 45433, AFFDL-TR-70-64 (1970).

[6] Bywaters, R. P. and Griffin, R. A., A Gas-Gap Thermal Switch for Cryogenic

Applications, Cryogenics 13, 344-349 (1973).

[7] Castles, S., private communication (1980).

81
SERVICEABLE REFRIGERATOR SYSTEM FOR SMALL SUPERCONDUCTING DEVICES

Ralph C. Longsworth

Air Products and Chemicals, Inc. APD Cryogenics

P.O. Box 2802

Allentown, PA 18105

The design of a refrigeration system for cooling small superconducting devices such as Josephson Technology computers which can be serviced

while the superconducting device remains in a cold operating state is

described. The refrigeration system incorporates new (passive) thermal

couplings which permit the Displex expander and JT circuit to be warmed

up, the expander serviced, and the JT circuit purged, while helium boils

off at a relatively low rate from the bath that is cooling the superconducting device.

When connected to the 1.4 L test pot, a warmup-purge-cooldown test was run. Helium was initially condensed in the pot by the refrigerator

at the rate of 0.3 L/hr then the refrigerator turned off. It took 7 hours to warm up and purge the refrigerator, followed by 3 hours to cool

back down to 4K. Just a few minutes before reaching 4K, helium stopped

venting and started to be drawn into the test pot indicating that

slightly more than 1.4 L of helium was used during the service cycle.

Key Words: Cryogenic refrigerator; liquid helium; Josephson technology;

serviceable; superconducting device; thermal coupling.

1. Introduction

As applications for small superconducting devices become commercial, there will be a need for closed-cycle refrigerator systems that permit such a device to be maintained in a cold operating state continuously, including periods when the refrigerator is off and

being serviced. At the 1977 meeting of this group, van der Hoeven and Anaker [1] described the refrigerator requirements for a Josephson technology data processing system

with particular emphasis on reliability and serviceability. At the same meeting Longsworth

[2] described means of refrigerating small superconducting devices, while providing isolation from vibration and electromagnetic "noise".

This paper describes the design of the refrigerator system developed for this application which incorporates new thermal couplings that permit the expander and JT circuit to be warmed up, the expander serviced, and the JT circuit purged, while helium boils off at

a relatively low rate from the bath that is cooling the superconducting device. A companion

paper to this reports on test results which show reasonably good agreement between calculated values and actual performance.

2. System Description

The basic refrigerator system is an APCI Model CS-308 two-stage Displex refrigeration system with a JT loop to provide refrigeration at /^4.2K. Refrigeration capacity

ratings are 2.5 watts at 4.5K plus 2 watts at < 20K and 10 watts at <77K. The original

JT circuit which was designed for cooling samples at 4.5K has been modified to permit

remote cooling of a liquid helium dewar using the first and second stage refrigeration to intercept heat leak and 4K refrigeration to recondense helium boil-off.

82.
The compressor is a three-stage oil -lubricated reciprocating unit that supplies gas at 17.7 atm and receives gas from the expander at 4.1 atm and from the JT loop at 0.8 atm.

In addition to the compressor, the standard CS-308 system includes a two-stage Displex

expander, the three JT heat exchangers and adsorbers, manually adjustable JT valve, a 4K

heat station, a hydrogen filled thermal switch connecting the second and 4K heat stations

(for cooldown), a radiation shield attached to the first stage, and the gas ballast tank.

Figure 1 is a piping schematic of the present system. The helium pot simulates the

helium dewar in the final system and serves the purpose of enabling the performance of the

thermal coupling system and thermal losses during servicing of the refrigerator to be

measured. The basic CS-308 system is thus modified by adding or substituting the following components:

a) thermal couplings 1 and 2 are added

b) a helium condenser replaces the standard 4K heat station

c) gas lines are added on the high pressure JT line to transport refrigeration to the separate thermal couplings and condenser

d) a bypass valve for cooldown is substituted for the thermal switch

e) a separate feed line for precooling and liquifying make-up helium is added

f) self-sealing gas couplings on the JT circuit are added to permit it to be isolated

g) purge lines on the JT circuit are added to permit it to be flushed

h) additional temperature sensors and heaters are added to permit control of rapid

warmup for service and monitoring of performance

i) an absolute pressure JT bypass pressure regulator for automatic control of the

liquid temperature is added

j) an experimental 1.4 L helium pot connected to a separate warm flange by a support

tube, having its own radiation shield, is added.

Refrigeration systems similar to this are presently being built for cooling superconducting magnets and spectroscopy samples in which the refrigerator and magnet or spectroscopy sample holder are in separate enclosures connected by a flexible hose containing the

gas lines.

An important design criteria that is adhered to is that the circulating refrigerant

helium, which must be kept free of contaminants, is never mixed with the helium that is in

contact with the superconducting device.

3. Refrigeration

Figure 2 shows the first and second stage capacity of the DE-208L Displex expander

before adding the JT loop. The width of the capacity curve reflects a small interdependence of the heat load- temperature relation for the two stages and manufacturing variations from unit to unit. When the JT loop is added there are additional heat loads

imposed, primarily heat exchanger losses on the first stage, and both heat exchanger

losses and enthalpy difference loading imposed on the second stage.

In order to determine the complex relationship between first, second, and JT heat

station heat loads versus temperatures, calculations were made of refrigeration losses in

the JT heat exchangers assuming the following matrix of operating conditions.

JT pressure at the compressor of 0.8 and 1 atm corresponding to 0.23 and 0.29 g/s

flow respectively

second stage temperatures of 16 and 20K

first stage temperatures of 50, 65 and 80K

83
84
I

u

<

u

UJ

I—a

z

8UJto

TEMPERATURE -K

Figure 2. CS-208L Refrigerator Capacity

85
Heat exchanger and enthalpy losses calculated at these conditions are subtracted from

the minimum capacity curve of Figure 2 to obtain the performance map shown in Figure 3. Figure 3 plots first and second stage refrigeration available versus first stage temperature for T2 = 16 and 20K and P,^ = 0.8 and 1 atm. Refrigeration produced at the 4K heat

station is also listed. Use of the curve is illustrated by assuming a 10 watt heat load

on the first stage, second stage temperature of 16K, and 1 atm JT return pressure. First

stage temperature is 55K, second stage capacity is 2.5 watts, and 2.3 watts is available

at about 4.3K. There is a small pressure drop loss in the low pressure side of the JT

loop, thus the saturation pressure at the 4K heat station is higher than the return pressure at the compressor.

30

I

40 50 60 70 80

FIRST STAGE TEMPERATURE - K i

I

i Figure 3. CS-308 Available Refrigeration of First and Second Stages i Versus First Stage Temperature. Capacity of the 4 K stage is tabulated

for second stage temperatures of 16 K and 20 K and return pressures of

0.8 and 1.0 atmospheres. The dashed lines at 56 K and 2.5 W refer to the example given in the text. i

86
Figure 4 plots the refrigeration available at the second and 4K heat stations as a function of second stage temperature for 0.8 and 1 atm JT return pressure assuming a 15 watt heat load on the first stage. This plot shows that if there is a 2 watt heat load

on the second stage and if the JT return pressure is 1 atm, then the second stage temperature is 16. OK and 2.3 watts is available at the 4K heat station.

These calculated capacities neglect conduction and radiation losses which can typically be kept relatively small by careful design.

4. Cool down

It takes about 2-1/2 hours to cool down the standard CS-308 refrigerator and about

3 hours to cool down the refrigerator and thermal couplings. If additional mass is in

good thermal contact with each of the heat stations then the first stage can cool 2.3 Kg

of copper per hour and the second stage can cool .9 Kg of copper per hour. For the

configuration shown in Figure 1, there is poor thermal coupling of the helium pot with the

heat stations, thus the rest of the system cools down much faster than the helium pot.

In a dewar with a large neck tube it is possible to provide means for the first and

second stage thermal baffles to cool down the helium pot by convection. Another option to cool the helium pot faster is to feed helium in through the liquefaction circuit and vent

it through the top of the neck tube. A small circulating pump could be used to eliminate

helium losses. Such an arrangement can be used as an alternate means to cool the leads

from a superconducting device such as a magnet.

5. Thermal Couplings

Figure 5 shows a cross section view of the first stage thermal coupling which transfers heat by convection from the base to the finned tube in the top which in turn is

cooled by flow from the JT loop. In the event that the refrigerator shuts down, the

thermal coupling stays cold because gas circulation in the JT loop stops. Helium will

typically boil off at a relatively low rate during this time, i.e., 0.1. to 0.5 L/hr.

The primary purpose of the convective thermal coupling is to enable the JT loop to be

warmed to room temperature and purged without significantly increasing the helium boil -off

rate. This is possible because the helium stratifies inside the thermal coupling when the

top is warmer than the bottom so heat is only transferred by conduction in the gas and

housing from top to bottom.

High heat transfer rates with modest temperature differences and relatively compact

design are achieved by separating the hot rising gas from the cold down-flowing gas and

locating extended surface heat exchangers in such a way as to promote convective circulation of the gas. Since the convective heat transfer rate is proportional to pressure, it

is best to keep the pressure at the high pressure side of the system. In other applications the thermal convective coupling can be used as a variable thermal conductance device

by adjusting the gas pressure.

The helium condenser shown in cross section in Figure 6 is designed with the JT stream cooling a finned tube in the top chamber which is separated from the main liquid

helium pot by standoff tubes. These standoff tubes are designed to have a low heat leak

to the liquid helium when the refrigerator is off and when the JT loop is warmed up to be

purged.

87
Figure 4. CS-308 Available Refrigeration of Second Stage and 4K Heat

Station Versus Second Stage Temperature with 15 watts on First Stage.

88
HELIUM PRESSURIZATION

PORT

Figure 5. Convective Thermal Coupling for First and Second Heat

Stations.

The JT loop gas coupling, the convective thermal coup! ings, and the helium condenser

are all (passive) thermal disconnect switches that serve to thermally isolate the refrige

ator from the liquid helium bath when the refrigerator shuts down and when it is being

serviced.

89
FINNED TUBE

HELIUM

CONDENSER

JT FLOW

HELIUM TO BE LIQUEFIED

.STNSTL TUBES

J Li.

FLANGE TO

HELIUM POT

Figure 6. Helium Condenser.

6. Service

very ?ew"case/nr"?^.\?H^"'?'

'"^'^^ closed-cycle coolers, it is seen that there are I!2 I ?f ^f'f developing in the vacuum chamber if all the joints are properly S brLk^^'c^ui if^'ftL''^'''-H ^'^ K^'*^^ sufficient to leave the'piping ?n 3 nd

^1 kI can ^^^""^^^ 3) the expander can be disassembled for service and, b) if the JT Iood be warmed and purged or possibly flushed.

^

90
Servicing a cold expander in less than an hour has been demonstrated. This was done

in the following sequence:

a) disconnect gas lines to the expander and let gas out

b) remove valve motor (the displacer stays in place)

c) screw a special tool with purge gas flowing through it into the top of the displacer

d) with purge gas flowing pull the displacer into a sealed plastic bag

e) warm displacer and service it while keeping a cover on the cylinder with a small

purge flow of helium

f) reassemble the expander in the reverse order

g) evacuate and purge the expander several times before reconnecting the gas lines.

The most common source of trouble with the JT loop is blockage due to freeze-out of

contaminants. Servicing of the JT loop thus usually consists of warming it to room temperature and purging or flushing out contaminants. This is done in the following sequence:

a) Disconnect the gas couplings to isolate the JT loop.

b) Warm the heat exchangers with a back flow of gas through the low pressure side of

the heat exchangers, venting out through the bypass line.

c) Simultaneously warm the refrigerator and adsorbers by turning on the heaters.

d) Warm the tops of the thermal couplings and helium condenser by flowing purge gas

in through the bypass line and out through the high pressure lines. Adjust the

heaters on adsorber 3 and the refrigerator while doing this to keep the heat ex- changers at room temperature.

e) Reconnect gas lines after warmup and purging is complete.

Liquid helium boiloff is minimized by keeping the tops of the thermal couplings cold

as long as possible - thus the refrigerator, heat exchangers, and adsorbers are warmed up first. Since the tops of the thermal couplings are relatively light they can be warmed

fairly quickly.

When this -procedure was first tried on the unit with 1.4 L of liquid in the test pot,

it took a long time to adjust the helium purge flow rate and set the heaters. The total

time from shutdown to return to 4.2K took 10 hours. Five hours were spent warming the

refrigerator to room temperature. With the refrigerator warm and the thermal couplings

cold, the boil-off rate was 0.13 liquid L/hr. The next 2.1 hours were spent warming the

tops of the thermal couplings and cold adsorber during which time the boil-off rate increased to 0.24 liquid L/hr. Three hours were then required to cool back down to 4.2K.

Helium stopped venting and started to be drawn into the feed line about 15 minutes

before the system reached 4.2K indicating that the entire warmup and purge consumed slightly more than 1.4 L of liquid helium. The unit then recondensed helium at a rate in excess

of 0.3 L/hr.

The warmup and purge period can be reduced from 7 hours to about 2.5 hours with some

small changes in instrumentation and practice.

91
7. Summary

i

1 Many thousands of small cryogenic refrigerators that operate at about lOK on the

Gifford McMahon and modified Solvay cycles are in use today. The most common small 4K

(systems) in use today use these lOK refrigerators as a base to achieve 4K by adding a JT loop. While it essentially doubles the complexity of the refrigerator, it is still possibl'i to achieve very high reliability, as described by Higa [3], if care is taken in the

construction, operation, and maintenance of the units. '

The refrigerator described in this paper extends the applications for this type of

refrigerator to cooling small superconducting devices such as Josephson technology data

processors, which in addition to high reliability, require an ability to service the

refrigerator while the superconducting device continues to operate. This has been accom- plished by means of thermal convective couplings which passively disconnect the refrigerato|

from a helium dewar when the refrigerator shuts down and permit the refrigerator circuit

to be warmed and purged.

Warmup, purging, and cooldown back to 4.2K has been demonstrated with a total loss of

about 1.4 L of liquid helium.

References

[1] van der Hoeven, B. J. C. Jr., and Anacker, W. "Refrigerator Requirements for

Potential Josephson Data Processing Systems", NBS Special Publication 508, p. 213,

April 1978.

[2] Longsworth, R. C, "Concepts for Cooling Small Superconducting Devices Using ClosedCycle Regenerative Refrigerators", NBS Special Publication 508, p. 45, April 1978.

[3] Riga, W. H. and Wiebe, E., "One Million Hours at 4.5 Kelvin", NBS Special Publication

508, p. 99, April 1978.

92
PERFORMANCE OF A 1 watt 4K CRYOSYSTEM SUITABLE FOR A SUPERCONDUCTING COMPUTER

E. B. Flint, L. C. Jenkins, R. W. Guernsey

IBM Thomas J. Watson Research Center Yorktown Heights, New York 10598

Superconducting computers of the future will require closed cycle refrigeration

systems capable of removing 0.5-10 watts at 4K. The refrigerators must be reliable and the interface to the cryostat must allow continued computer operation during

refrigerator shutdown. Such a system has been designed by R. Longsworth and one

unit has been built by Air Products. This cryocooler is being operated now in a test configuration that allows detailed study and analysis of the design. The results of

these studies, including cool-down times, refrigeration capacity, and interface per- formance will be discussed and compared with the design calculations.

Key words: Convective heat transfer; cryocooler; heat transfer by evaporating and

condensing helium; helium refrigerator; superconducting computer; thermal coupUngs; thermal isolation.

1 . Introduction

1 . 1 Desired cryosystem characteristics

In the future, one may be able to buy an extremely high-speed computer whose logic and memory chips operate

at 4.2K. This might be packaged in a cube 9 cm on a side and would be immersed in liquid helium where it would

dissipate about lOW of heat [1]. At this time, it is of interest to evaluate cryosystems that might be appropriate for such a machine. A general view of the requirements, as anticipated at this time, is given elsewhere in these proceedings.

Medium to large size superconducting computers are likely to need from 0.5W to lOW of refrigeration at 4.2K. The cryosystem should provide continued computer operation in the case of refrigerator shut down, and it also should allow one to service the computer without disturbing the refrigerator. A semi-annual service interval and a mean time between failure in excess of two years would be desirable. High reliability and fool proof service procedures are paramount. This paper documents the performance of a refrigerator and thermal interface designed

to support a IW, medium sized superconducting computer.

1.2 The interface concept

A study was done by R. Longsworth of Air Products, in which he developed a design for a complete cryogenic

system using existing refrigeration technology which meets the requirements of a superconducting computer. The

design he developed uses the Air Products CS-308, 4K refrigerator to which was added a thermal interface which

I couples the dewar and bath to the refrigerator. A schematic of the system is shown in Fig. 1. The interface provides good thermal coupling between the refrigerator and the dewar structure and processor during normal

operation and good isolation when the refrigerator is warm for servicing. A companion paper by R. Longsworth [2] describes the design in some detail. Following the study. Air Products [3] built the system in a test configuration

that now is undergoing extensive tests at IBM. While this work is far from complete (for example, there are few

reliability data) there are data characterizing the refrigerator capacity and interface performance. In this paper we

93
PURGE

PORTS

REFRIGERATOR

260 PSIA

SUPPLY

60 PSIA

RETURN!

JT RET.

BY- PASS

VALVE

300K

GAS

DISPLACER

DRIVE MOTOR a

VALVES

COMPUTER

®

; 1. VALVE ~\

, IMSJ]-

^4K LIQUID ' rt -_ JT GAS RETURN -J

VACUUM

RAOIATION- SHIELD

CONDENSERLIQUID HELIUM BATH

AT 4.2

K

-@-

-SEMICONDUCTOR

INTERFACE

-GATE VALVE

I/O CABLES

SUPPERC0NDUCTIN6

MAGNETIC SHIELD

-COMPUTER LOGIC a MEMORY CARDS

Figure 1: Shown is a schematic representation of a superconducting

computer cryosystem with a large liquid helium reservoir

for thermal ballast. The cryosystem under test has a 1.4 liter pot and includes thermometers T1-T8 and heaters

H1-H6

.

94
describe the cryogenic system and instrumentation used in the experiments, the performance data ^ and the areas where we believe the data indicate more work, both documentation and engineering, is required.

2. Description of the cryogenic system

In this section we describe the system including the refrigerator, the thermal interfacej and the instrumentation used in the experiments. Figure 1 shows the refrigerator and interface with a dewar which might be suitable for a

processor. The large liquid helium bath provides thermal ballast during periods when the refrigerator is shut down,

i In the test configuration a 1.4 liter hehum pot replaces the bath shown.

! The 4K refrigerator includes the 2 stage expander which provides refrigeration at 60K and 15K and a Joule-Thompson (JT) loop which produces the refrigeration at 4K. There is a by-pass line around the low pressure

side of the third JT heat exchanger which is used for the cool-down from room temperature. The JT valve can be

adjusted from the top of the cryostat so as to compensate for different load conditions.

The thermal interface consists of two gas thermal diodes and a condenser. The refrigeration available at the

i first and second expanders is delivered to the tops of the diodes via the JT gas stream. The diodes conduct by gas

j convection from bottom to top. When the temperature gradient is reversed the gas stratifies, providing relatively

good thermal isolation. The condenser is similar to the diodes except that it is a two phase device. In normal

I operation (top cold) the boil-off gas from the bath is condensed on fins cooled by the JT stream and then runs back

; into the bath. Thermal isolation is, once again, provided when the top is warm because the helium gas in the support tubes stratifies.

The system performance has been studied using instrumentation which includes thermometers and heaters at each of the refrigerator stages. At the first and second stages silicon diode thermometers (numbers: Tl, T2, T5, T6,

! T7, T8) are used to monitor the refrigerator temperature, the heat station temperature at the base of the thermal

j diode and the temperature of JT gas as it returns from the diode. The thermometer, T2, was calibrated between 4K

and lOOK. All other silicon diode thermometers have four point caHbrations (4.2K, 20. 4K, 77.4K and 273. 2K)

which were used to scale the T2 temperature vs. voltage curve to obtain temperatures from the measured voltages.

The relative accuracy of the thermometers is probably less than IK at 70K and less than O.IK at 20K. Thermometers T3 and T4 are calibrated germanium resistance thermometers located inside the JT stream. The manufacturer quoted accuracy of these thermometers is 0.005K. The temperature of the bath was measured by monitoring the vapor pressure and using the 1958 helium temperature scale.

There are six heaters (HI < 60 watts, all others < 20 watts) which were used to simulate heat loads and to warm the refrigerator for purging the JT loop and servicing the refrigerator.

3. Refrigerator performance

3.1 Capacity

The refrigerator was delivered with the thermal interface, test pot, and instrumentation in place. In order to determine the total cooling capacity at each stage, one must know the parasitic heat loads that result from these

I structures. We begin with a discussion of the methods used to measure these. At the first and second stages heat

leaks are intercepted at the refrigerator stage and the coupling station. Those for the two couplings were determined

from the temperature increase in the JT gas stream and the JT flow rate, and these values were confirmed by

calculations of heat leak from the support structures and from radiation. An additional check was made by

extrapolating the heat transport characteristic, Q(AT), to AT = 0 for the couplings. Q is the heat transported across

the coupling and AT is the temperature difference between the JT gas (either entering or leaving the coupling) and

base of the coupling. In the case of the first stage data, the parasitic loads are sufficiently large and the Q(AT)

characteristic sufficiently non-linear that it is possible to conclude only that there were no inconsistencies in the

data. For the second stage coupling the loads (parasitic and applied) are smaller and the three determinations of the

I

parasitic load were in reasonable agreement.

I The parasitic load to the JT stream from the test pot was measured by noting the boil-off rate with the JT and

I the pot thermally decoupled. The result is in reasonable agreement with the calculated value. The parasitic load to

I the pot as indicated by the Q(AT) characteristic for the condenser extrapolated to AT = 0 was substantially higher

and the difference is not understood. Offsets of ~ 0.035K and ~ 0.008K in the JT thermometers up and down

stream of the condenser relative to the bath temperature would be required to explain the discrepancy. The parasitic loads at the three refrigerator stations are given in Table 1.95
A complete map of the refrigeration capacity has not yet been made. The capacity at 4K, however, has been

measured with the first and second stages loaded as they would be in an actual computer cryostat. Table 1 shows

the results of these tests. With 9.9 watts applied to the first stage (22W total) and 0.5 watts to the second stage (IW total) the minimum temperature in the pot was determined with appHed loads of LOW and 1.5W to the bath. At 1.03W total dissipation in the bath the minimum temperature, T^j^ was 3.82K and at 1.53W total dissipation T^in was 4.24K. A summary of data is given in Table 1.

The refrigeration at 4.2K is 25% less than anticipated but the refrigeration at the other stages is near the design

values.

Table 1

Station

Parasitic Load (W) Applied

Heat Load

Total

Heat Load (W)

Stage

Temp.

Temperatures at

Couplings and Pot

Refrigerator • Coupl.2

1st 2nd JT

5.5 0.206.5 0.24

0.03

9.9

0.5 . 1.0

21.9 0.94

1.03

59

14

3.77 (est)3

86K (est) 15.5 (est) 3.82

(.57 atm at compressor)

1st 5.5 6.5 9.9 21.9 57 84K (est) 2nd 0.2 0.24 0.5 0. 94 16.3 17.8 (est) JT 0 0.03 1.5 1.53 4.18 (est) 4.24

(0.92 atm at compressor)

Notes:

1) heat load in this column are calculated; 2) heat loads in this column are measured and checked with calculation;

3) est. = estimated temperature. Temperature is determined from one measured temperature and the coupling heat

transfer characteristic.

3.2 ReUability

It is important that a superconducting computer cryosystem be highly reliable and a maintenance interval of 6 months or longer is desired. Reliability studies will become a major part of our test program. When the refrigerator was delivered the compressor had run 1375 hours, and it has run an additional 1685 hours at IBM, bringing the total operation time to 3060 hours. Although this is only one third the scheduled maintenance interval, the refrigerator has been shut down and warmed to room temperature several times to correct problems, primarily with the instrumentation. Once the initial debug and performance mapping is completed, long term reliability testing will begin.

4. Interface performance

The thermal interface between the refrigerator and dewar is a key element of the cryosystem. It must deliver the refrigeration efficiently in normal operation and provide good thermal isolation during periods when the

refrigerator is warm for servicing. It also must provide for liquefaction of make-up heUum. The thermal contact provided during cool-down from room temperature must be good so that the cooUng time required is reasonable. In

this section we present the test results for the thermal interface system and the JT temperature regulator.

96
4.1 Thermal coupling heat transfer characteristics

The heat transfer characteristic, Q(AT), for the first and second stage thermal diodes was measured by heating

the base of the diode while monitoring the temperatures of the ingoing and outgoing JT gas and the diode base. For

the second stage diode, thermometers T2, T7 and T8 were used to measure the temperatures. The results are shown

in Fig. 2 where the heat transported through the diode is plotted as a function of the temperature drop across the

diode. The circled points are taken fromR. Longsworth's design calculations and their locations below the measured

curves indicate that the diode performs somewhat better than anticipated. The two sets of data for this diode were

fit to a straight line which gave slopes (diode conductance) of 0.81 W/K and 1.85 W/K and intercepts of -0.24W

and -0.3 IW. The intercepts should be the parasitic load and in fact are in approximate agreement with the

calculated value and also with that determined from the measured enthalpy increase in the JT gas.

Second Stage Diode

ID

T - T diode JT out

T - T diode JT in

AT (K)

Figure 2: Heat transfer characteristic for the second

stage diode.

97
Similar measurements and analysis were made for the first stage diode. There is some doubt as to the absolute accuracy of T6 which measures the diode temperature but it can be used to measure the temperature changes at the base of the diode. It therefore was not possible to obtain parasitic losses from these data. In addition, the heat

transfer characteristic for this diode is non-linear which makes the analysis more difficult. The slopes obtained from

fitting the data through Q = 15W total load to a straight line give a measure of the diode's effective thermal

conductivity. For AT = T^io^^ - Tjj the conductivity obtained is 0.69 W/K and for AT = T^io^^ - Tjj the conductivity is 1.11 W/K. If instead one fits the same data to a second order polynomial to allow for the curvature, the corresponding effective conductivities for 15W through the diode are 0.75 W/K and 1.22 W/K. The numbers

obtained fromR. Longsworth's design calculations are 0.75 W/K and 1.9 W/K. One can conclude that the coupling

performs substantially as designed.

The helium condenser used to couple the bath to the JT stream was tested similarly at several bath temperatures

between 3.6K and 4.5K. The results for a nominal bath temperature of 4. OK are shown in Fig. 3. The squares

represent the temperature difference between the bath and the JT loop upstream of the condenser, and the circles give the difference between the bath and the JT loop downstream of the condenser. R. Longsworth has calculated a temperature drop of 0.1 58K for 2 watts through the condenser. This is equivalent to a thermal conductivity of 12.7 watts/K. The values obtained from fits of the data to a straight line typically give 18-26 W/K with no discernable dependence on the operation temperature in the range of 3.6K < T^ath < 4.5K. The condenser is at least 50% I more effective than anticipated.

j

4.2 System performance of the interface

The performance of the thermal interface in normal operation is shown in Table 1. The temperatures marked

"estimated" have been determined from the heat transfer characteristic for the appropriate coupling. As the temperatures in the table indicate, the refrigeration is very effectively coupled to the heat stations and the bath. A test of the thermal isolation provided by interface when the refrigerator is warm was conducted as follows. With the refrigerator in its normal operating mode the 1.4 liter test pot was filled. The refrigerator was then turned

off and warmed to room temperature with the heaters on the two expander stages, the heater on the third adsorber,

and a flow of 6 SL/M of helium gas through the low pressure side of the JT loop. The average boil-off during the

2.5 hours required to warm the refrigerator to room temperature was 0.055 liters/hr. At 2.15 hours 2.4 SL/M of

the He gas flow in the JT was directed out the high pressure side. This warmed the tops of the two coupHngs and

the condenser and increased the boil-off from the pot to 0.15 Hter/hr. The refrigerator was restarted at the end of

the purge, 3.15 hours after it had been shut down. The total helium loss during this period was 0.25 liters. After

the refrigerator was started the boil-off increased to 0.24 liter/hr. It is likely that this increase indicates that the tops of the couplings had not been fully warmed to room temperature and a longer purge is required. Assuming

normal operation 3 hours after the restart of the refrigerator, as demonstrated by Air Products, and a 0.20 liter/hr average boil-off during this time the total helium loss would be approximately 0.8-1.0 liters. Some adjustments in the purge procedure may make it possible to reduce the time and/or helium loss further. It is worthwhile to note

that the 1 liter of liquid which is lost due to the cryogenic structures and the reverse heat leak of the thermal

couplings is small compared to 8.4 liters which would be boiled by a 1 watt processor whose operation is maintained

during this time.

• 4.3 Liquefaction

The liquid lost during a refrigerator shut down must be replaced by the liquefaction of He gas supplied from room temperature. . This is done by admitting the gas to the condenser via a tube that is thermally sunk at the first and second stages of the refrigerator. The liquefaction rate has been measured with and without additional heat

loads applied to the three rpfrigerator stations. With no applied loads (parasitic losses only) the Hquefaction rate is 6.7 liter/day. With simultaneous loads of lOW, 0.56W, and 0.23W applied to the first and second stages of the

refrigerator and the helium bath, respectively^ the liquefaction rate dropped to 3.9 Uter/day. Assuming ideal heat exchange at the first and second stages the 6.7 liter/day liquefaction rate represents a load of 11. 8W, 2.3W, and

1.05W to the three stages of the refrigerator. The 1.05W load at 4.2K is roughly 50% less than the measured

refrigeration capacity at this temperature and suggests that the heat exchange at the second stage is not complete. It may be possible to increase the liquefaction rate through better exchange at the first and second stages and/or

thermally sinking the condensing line to the third JT heat exchanger. It is not essential, however, in this application

that the liquefaction rate be optimized.

98
ID

CVJ

CD

LU ^ o

o

1 1 1 r

COM)ENSER

I r

Tboth = 4. OK

± ±

0.02 0.04

AT

0.06

(K)

0.08 0.10

Figure 3: Heat transfer characteristic for the condenser. The

squares are determined from the temperature difference

between the bath and the JT temperature upstream of the

condenser and the circles from the JT temperature downstream of the condenser.

99
4.4 Cool-down from room temperature

The time required to cool the helium test pot to 4.2K starting with the entire system at room temperature is approximately 10 hours. The first and second stages cool more quickly, nearing their operating temperatures in 3-4

hours. The difficulty is that the condenser is designed as a 2-phase device and provides only a very weak thermal

link at higher temperatures. One could achieve a more rapid cool-down with the available refrigeration if, for example, a hydrogen gas thermal coupling were added betweeen the pot and the second stage. Another alternative would be to modify the condenser geometry to enhance gas convection. A model calculation, which assumes good

contact between the He gas and the pot, indicates that such a modification could reduce the time required to cool

the pot to 6 hours. The changes in the geometry of the condenser would not compromise its normal operation or the

reverse heat leak. It is expected that the therml contact between the gas and pot would have to be improved with

fins.

4.5 JT temperature regulator

A JT pressure/temperature regulator was supplied with the refrigerator. A sense pressure, referenced to vacuum, causes an increase or decrease in gas flow through the regulator from the expander return to the JT return which regulates the JT return pressure. In our tests the sense pressure was the saturated vapor pressure in the helium pot. Figure 4 shows the results of the tests. The horizontal axis is the heat applied to the liquid in the test pot and vertical axis shows the resulting temperature increase. The upper curve was taken with no regulation and

represents the ability of the refrigerator to absorb the increased load. A linear fit gives a response characteristic of 0.039K/W. The lower curve gives a regulated response of 0.013K/W. This represents roughly a factor of 3 improvement.

There is a mechanical hysteresis in the regulator of about 20 mm Hg which corresponds to a temperature

difference of 0.028K at 4.2K. In fact, it was necessary to tap the regulator continuously to obtain the data in the

figure. Air Products had measured a substantially smaller hysteresis (8 mm Hg) prior to delivery. The cause of the

hysteresis has yet to be studied.

The performance of the regulator has not yet been measured at lower bath temperatures. Since both the sense and the feedback response generated by the regulator depend on the slope of the saturated vapor pressure curve, dP/dT 1 53,, the gain of the system may be essentially unchanged at lower temperatues. Mechanical hysteretic

effects, however, are likely to be more pronounced.

Summary

Tests of an Air Products refrigerator and thermal interface designed to support a superconducting computer

have given the following results. In normal operation the interface provides good thermal contact between the dewar structure and the bath with a minimal loss of refrigeration. The liquid bath can be maintained automatically

between 4.20 and 4.22K for applied heat loads varying from 0 to 1.6W. The interface also automatically isolates the dewar and bath in the case of refrigerator shut down.

The procedure and/or instrumentation used to warm the refrigerator for servicing ^nd purging should be

improved to reduce the time and liquid lost. The time required to cool the system can be and should be reduced by

the addition of a thermal switch between the second stage and the helium pot. This will require a change in the hardware. The long-term reliability of the compressor package and refrigerator has yet to be studied.

100
JT Temperature Regulator

1 1 1 1 \ 1 1 r

I I I I \ I I I i_l

0 0.4 0.8 1.2 1.6

POWER (W)

Figure 4: Increase in bath temperature, AT, as a function of

the power applied to the bath with and without the

JT regulator.

101
Acknowledgement

We would like to acknowledge the assistance given by J. Stasiak with the data acquisition.

References

[1] Anacker, W., et al, "Josephson Computer Technology," IBM J. Res. Dev., 24 [2], 107-264 (March 1980).

[2] Longsworth, R., "Serviceable Refrigerator System for a Small Superconducting Device," Proceedings of the Conference on Refrigeration for Cryogenic Sensors and Electronic Systems, NBS, Boulder, CO, Oct. 6 and 7, 1980.

[3] Air Products and Chemicals, Inc., APD Cryogenics, AUentown, PA, 18103.

102
DEVELOPMENTS TOWARD ACHIEVEMENT OF A

3-5 YEAR LIFETIME STIRLING CYCLE REFRIGERATOR

FOR SPACE APPLICATIONS

Max G. Gasser and

Allan Sherman

Cryogenics, Propulsion and Fluid Systems Branch

Space Technology Division Goddard Space Flight Center

Greenbelt, Maryland 20771

and

WilMam Beale

Sunpower, Inc. Athens, Ohio

A major effort is underway at the NASA/Goddard Space Flight Center to develop a long lifetime Stirling cycle cryogenic cooler for spaceflight. The initial objective is a 3-5 year, 5 watts at 65°K, cooler to be followed by a 12°K cooler. The

previous paper, "Progress on the Development of a 3-5 Year Lifetime Stirling Cycle

Refrigerator for Space," presented at the 1979 Cryogenic Engineering Conference,

outlined the approach [ 1 ] . During the past year, the design of a refrigerator with linear magnetic bearings has changed. The displacer rod no longer goes through the piston,

resulting in a machine that is comprised of a compressor section and an expansion section bolted together. Each section has its own magnetic bearings, linear motor, displacement

transducer (LVDT), and displacement volume. The phase angle and displacements of

the two moving parts are electronically controlled, utilizing the signals from the LVDT's.

Delivery of the engineering model magnetic bearing cooler is scheduled for March, 1 98 1

.

Work on the engineering model gas bearing cooler previously reported has ceased. However, an advanced concept for a gas bearing cooler is being studied.

Preliminary results indicate that the normal gas flow in a free displacer Stirling

cycle cooler can be utiUzed to spin the displacer at a speed sufficient to form a self-acting

hydrodynamic gas bearing with a clearance seal. The periodic gas flow from the cold end

through the generator to the ambient end makes intermittent contact with a turbine

located at the ambient end of the displacer. The induced spin is unidirectional.

Key words: Cryogenics; space application; long lifetime; Stirling cycle; linear magnetic

suspension; clearance seals; hydrodynamic gas bearing.

103
1. Introduction

A wide variety of spacecraft instruments require cryogenic cooling to accomplish mission objectives. There

are many cooler systems which have been developed or are being developed to satisfy these needs [2] . A particular

critical area is the requirement for a system capable of providing high cooling-load (5-10 watts) for 3-5 year missions. A major effort is, therefore, underway to develop a long-lifetime Stirling cycle cryogenic cooler for spacecraft.

The basic problem is to develop a closed cycle machine with moving parts that will reliably run for billions of

cycles while unattended. Past attempts to solve this problem have used liquid or semi-liquid lubricants, dry lubricants, and "hard-on-hard" bearings with little or no lubricants [3] . Various seal designs were also tried. To-date,

these approaches have not yielded a long-lifetime spacebome cooler system.

The new approach initiated by NASA/Goddard Space Flight Center hopefully eliminates the problem-causing

features of the past. The key elements of the approach are: (1) The reciprocating components are driven directly

with linear motors. (2) While operating, there is no contact between the moving components and the machine housing and motor. The noncontact operation can be achieved by either magnetic or gas bearings and clearance seals.

The magnetic approach is presently being developed into an engineering model cooler for test and evaluation. The fluidic approach is still in the technology phase.

2. Cooler Utilizing Magnetic Bearings

2.1 System description

Figure 1 is the schematic of the North American Philips single expansion cryogenic cooler with linear magnetic suspension. There are three subassemblies: (1) the compressor section, (2) the expansion section, and (3) the balancing device. The displacer rod does not go through the piston, and at no time do the piston and displacer

occupy the same space. Each subassembly has its own linear motor, bearings, and displacement transducer. The

phase angle can be changed by the input signals to the motor. The Unear motions will be sinusoidal. The displacer

and piston are free to rotate, but are not intended to do so. There are no springs on the displacer; it is driven in both

directions. The piston operates with a gas spring. There are two pressure transducers (not shown in the figures): (1) to measure the static pressure, and (2) to measure the dynamic pressure cycles. Temperature transducers (not shown)

are mounted on the exterior of the pressure shell at the cold end and the ambient areas. A small electric heater (not

shown) is located at the cold end to simulate a load.

Figure 2 shows the compressor section at mid-stroke. The motor is the moving magnet type with the magnets

and coils sealed in metal cans to prevent outgassing products from reaching the working fluid. The piston is hollow

and vented to the motor housing. The close clearance (0.0025 cm) between the piston and its cylinder forms the noncontact seal with the piston centered in the cyUnder by the magnetic bearings in the attraction mode.

The bearings are comprised of four nickel/iron poles acting on a vanadium Permendur shaft. Two eddy curo rent sensors, positioned 90 apart around the circumference, measure the gap. The error signal from these sensors is then used to adjust the current in the magnets to prevent contact between the moving member and the cylinder wall. A redundant set of eddy current sensors is also incorporated in the design.

Figure 3 shows the expansion section at mid-stroke. The cold end has a heavy copper cap, which is connected

to the ambient flange by a thin titanium tube for low thermal conductivity. The cold heat exchanger is the gap be- tween the displacer and the copper cap. The cold end of the displacer is hollow and contains a vacuum. Regeneration

is by the phosphorbronze screen matrix and also in the gap between the displacer and the cylinder. (Again, in the seal area, this gap is 0.0025 cm.) The ambient temperature flange supports a vacuum dewar (not shown) around the cold

end. Starting at the cold end, the helium working fluid flows through a central passage in the displacer in the location of the magnetic bearing. This is the part of the displacer which constitutes the clearance seal. The working fluid flows outward through radial passages, over the linear motor, between the rear magnetic bearing, around the LVDT,

and into the piston compression volume. The displacer motor is the moving magnet type with the coils and magnets

sealed in metal cans. The displacer bearings operate in exactly the same way as that described for the piston.

104
EXPANSION VOLUME

REGENERATOR

IN DISPLACER

]300°K FLANGE

DISPLACER

LINEAR MOTOR

(MOVING MAGNET)

WASTE HEAT REJECTION

DISPLACER LINEAR

POSITION SENSOR (LVDT|

COMPRESSION VOLUME

HELIUM

FILL AND VENT VALVE

COMPRESSOR LINEAR MOTOR

(MOVING MAGNET)

COMPRESSOR LINEAR

POSITION SENSOR (LVDT)

Figure 1 . Schematic of the North American Philips single expansion cryogenic cooler with

linear magnetic suspension.

105
Figure 2. Philips Laboratories linear compression section of the Stirling cycle cooler shown in Figure 1,

All of the waste heat is removed by cooling jackets around the outside of both the piston and displacer housings.

Figures 4 and 5 show the passive counter-mass on cantilever support springs, which is one form of balancing

device. The active balancing system, which utilizes accelerometers, a linear motor (moving iron type), and magnetic

bearings, is described in reference [4] . The balancing device is bolted to the ambient end of the refrigerator. There is

COOLING JACKET FOR WASTE HEAT REMOVAL AROUND THIS SECTION

EXPANSION SPACE —

HELIUM GAS TYPE FLOW PASSAGES

Figure 3. Phihps Laboratories expansion section of the Stirhng cycle cooler shown in Figure 1. 106
LEAF SPRINGS

MASS

COIL SPRING

HOUSING

Figure 4. Cross-section of the counter-mass passive balance system for Figure 1

.

no communication of the working fluid to the balancing device. The balancing could also be accompUshed by two

refrigerators back-to-back and operating in counter-phase.

The passive spring-mass system has been used on the Philips MC-80 (Einhoven) cryogenic refrigerator and is tuned to vibrate counter-phase to the piston.

HOUSING

LEAF SPRING

MASS

Figure 5 . End view of the counter-mass passive balance system for Figure 1

.

107
2.2 Performance characteristics

The refrigerator is designed to produce 5 watts of cooling at 65°K when operating at a resonant speed of 1700 \ strokes/minute. The helium charge pressure is 16 ATM. The refrigerator will be one meter long with the largest diameter about 20 cm. The expected lifetime is 3-5 years with either continuous operation or 1000 starts. Noncontact between moving and static components can be achieved for startup, steady operation, and shutdown because the bearings can be activated independent of the linear motors.

The piston and displacer can be locked in place for launch by (1) energizing adjacent magnetic coils or (2) using the motor to hold the moving part in one of its extreme positions or (3) a combination of the two.

2.3 Component testing

Figure 6 shows a test setup to verify magnetic suspension in a refrigerator configuration. A shaker with a

linear DC motor was connected to the bearing shaft via an impedance head (an integrated accelerometer and force

IMPEDANCE HEAD

(INTEGRATED ACCELERMETER

AND FORCE J5AN^DUCER)

Lil)IEAR MAGNETIC BEARING

TEST DEVICE

NTLY-NE

POSITION SENSORS

FREQUENCIES BETWEEN

5Hz AND IK Hz

• 1 MIL RADIAL CLEARANCE

• DC STIFFNESS 16,000 LB/IN

• AC STIFFNESS 13,000 LB/IN

AT 25 Hz

Figure 6. Test setup to verify magnetic suspension in a refrigerator configuration.

108
transducer). Two Bently-Nevada position sensors were mounted to measure the displacement of the shaft center-ofgravity and the shaker armature. An accelerometer was mounted to monitor the deflection of the bearing housing.

The coil currents were observed with oscilloscope probes. Readings of the amphtude and phase of all quantities were

taken with a gain/phase meter via a narrow band filter for frequencies between 5 Hz and 1 kHz. The shaker was set to produce a DC force to counteract the effects of gravity and oscillate about the center position since this is the minimum stiffness condition.

The preferred bearing configuration based on brazing tests and magnetic hysteresis measurements of ring

samples is a combination of nickel/iron pole pieces and a vanadium Permendur armature with a 0.0025 cm radial clearance. Each side of the prototype of this configuration was found to have a DC stiffness of 16,000 lb/in. (28,000

Newtons/cm) and an AC sfiffness at 25 Hz of 13,000 lb/in. (22,800 New^ons/cm),

Figure 7 shows the piston motor armature with the magnet segments mounted, but before enclosure in a metal can. The motor will be tested on a special test rig before assembly into the refrigerator.

For comparison purposes. Figure 8 shows parts of the displacer motor. No test is planned for it before assembly into the refrigerator.

Figure 7. Piston motor armature under construction with magnetic segments mounted.

109
DISPLACER MOTOR STATOR

f ^ WITH COILS WOUND

Figure 8. Displacer motor parts.

3. Self-Actuated, Spinning Gas Bearing Feasibility Study

Sunpower, Inc. made a study to determine the feasibility of spinning the displacer of a Stirling cycle cryogenic

cooler by the interaction of the gas flow with a turbine mounted on the displacer. The turbine encounters the flow only at the outermost part of the displacer stroke, where gas flow is at its maximum rate, in such a way as to give a one-directional torque to the displacer.

The hydrodynamic bearing requirements for noncontact operation of the displacer could be met by such a turbine action, and the method indicates minimal penalty in flow losses. The bearing should be stable, and its fits and tolerance requirements should be well within manufacturing capability.

A design of a demonstration model has been made which is capable of showing hydrodynamic bearing operation, self-spinning and zero mechanical contact. no
The concept of the turbine spun, hydrodynamic displacer system having a bearing and turbine design compatible with a single-stage Stirling cryocooler of 5 watt heat lifting capacity at 65° K is described in the following paragraphs.

3.1 Hydrodynamic bearing system concept

Figure 9 shows a turbine spun displacer and an electrically driven piston which may either be spun

or supported by hydrostatic bearings or magnetic bearings. This is the concept which we would like to develop into a proof-of-principle device.

3 .2 Proposed spin bearing demonstration model

Figure 10 shows the proposed proof-of-principle device.

The displacer has a mass of approximately 50 grams, and this can be sustained in a 1.5 g field with a slow

speed gas bearing having a bearing gap of about 7 micrometers — a value typical of existing Sunpower seal clearances.

For purposes of turbine design, it would be desirable to spin the bearing more rapidly than the maximum

stable 10 radians/second, but this, in theory at least, can lead to plain bearing instability (whirl) under low loading as would be in the case in space. The use of herringbone rather than plain bearing surfaces gives promise of elimination of whirl instability at low load and high speed.

The proposed design will be first tested with a plain bearing. If this proves unsatisfactory, a herringbone will be chemically etched on the plain bearing surface (the depth of the herringbone groove is only about 20 micrometers).

COLD

RECIPROCATING

ROTATING MOTOR

Figure 9. Sunpower, Inc. long life cooler concept.

Ill
Figure 1 0. Sunpower, Inc. proposed spin bearing demonstration model.

3.2.1 Turbine design

The basic problem of the turbine spun bearing system is that while the spinning bearing requires very low

torque to maintain its load carrying capacity, its static friction starting torque is many times greater, at least as indicated by elementary static friction estimates. The high starting torque requires a high tangential gas velocity

impinging on the turbine, which would impose an excessively high running speed, possibly inducing large out of bal- ance forces. But, this is by no means an intractable problem. Many solutions are possible of which the least complex

might be a centrifugal brake (fan) on the displacer. However, one is reminded of the many occurrences in the past in which great effort was expended to solve on paper what in the lab was found to be a nonexistent problem. Just what

the startup torque requirement is in a reciprocating spinning seal with a pressure difference across it is apparently not known, and it may be very much lower than the static friction torque.

The simplest approach is a plain bearing and a brute force turbine capable of overcoming full sliding friction and accelerating the bearing to a high speed. Combined with the strong turbine, a geometrically similar fan brake is included, which is intended to limit displacer angular velocity to one tenth of the theoretical velocity capability of the

turbine. Both the turbine and brake were designed for ease of experimental variation, with the assumption that some

modification would inevitably be required.

The turbine was designed such that it is energized only about 1/8 of the cycle time. Its loss is in the range of

tenths of watts at most, and little effect on the thermodynamic performance of the cooler is expected.

3.2.2 Instrumentation

The purpose of the demonstration model is to show hydrodynamic lubrication without contact as the dis- placer cycles in a manner compatible with operational cryocooler performance. In order to show this, instrumentation must be provided which shows presence or absence of contact. 112
The system uses two proximity sensors at 90° separation in two separate planes perpendicular to the displacer

axis. These show a circle on an oscilloscope within which a dot representing the center line of the displacer runs, with

contact represented by the dot reaching the inner edge of the circle. This display would show not only contact, but

also eccentricity and whirl, and would be very useful in diagnostics.

3.2.3 Characteristics of proposed designs

The characteristics of the demonstration model are as follows:

1 3 bar rbpHiim^

Frequency 25 Hz

Displacer diameter 20 mm

Bearing length 40 mm

Mass 75 grams (for demonstration purposes)

Rod diameter 5.3 mm

Stroke 8 mm

Rotation rate 300 radians/second (3500 RPM max)

Piston diameter 52 mm

Stroke 8 mm

Drive motor — Type Dayton PM 42143, 12 volt — Power 100 watts — Speed 1750 RPM

The displacer is rotated by a turbine, which has gas flow for only 1/8 of the total stroke. During the remainder of the displacer motion, the turbine stators are short-circuited by the normal warm heat exchanger

ports. Volumes are sized so that the mass flow through the ports is the same as would be present if the cryocooler

temperature existed. This was done by enlarging the volume of the demonstration model cold end by the temperature ratio, so that it represents a larger mass flow source and sink, as does the cold end of the cryocooler.

The displacer has a fan brake which will initially be blocked. If experiment shows problems with overspeed,

the dam will be progressively removed, and the brake brought into greater effectiveness until the desired displacer

rotating speed is met. At full effectiveness, the brake is expected to limit rotating speed below 300 radians per second.

Materials are anodized aluminum and chrome-plated aluminum ground and honed to true roundness and

straightness. These have given good service in past applications to small Stirling heat engines.

3.3 Sequence of operations of the spin mechanism

Figure 1 1 shows the sequence of operations which causes the displacer to spin as it reciprocates in the normal

mode of the free displacer driven by fluidic forces.

The displacer motion leads the piston by 90°, that is to say, at a time when the displacer is at the midpoint of

its motion, going out (to the right), the piston is at the innermost point of its motion, just starting its outward motion

(out).

When the displacer is at its outermost point, where the turbine has come into register with the turbine stator, the piston is at its midpoint and moving with maximum velocity outward. At this point, the maximum flow rate of

gas occurs from the cold expansion space, through the regenerator, through the stator and turbine blades (since the warm port is blocked by the displacer), and then through the connecting tube to the piston space.

It is useful to note that this maximum mass flow occurs at a point (about 190°) where the displacer is almost

stopped, having just turned around and started in again. The flow occurs as a result of the influence of the motion of

the piston, which is at its maximum rate. Cold gas in the expansion space moves through the regenerator, warming

and expanding, then moving through the turbine toward the piston.

113
GAS MASS FLOW RATE FROM WARM HEAT EXCHANGER PORT VS. DISPLACER POSITION WARM PORT

.4ji

PISTON WARM

COMPRESSION

SPACE. CONNECTING

TUBE — IN COLD

EXPANSION SPACE

180' DISPLACER

0° I 360

h. ' .1 IN

DISPLACER AT 200 AS ROTOR PORT FROM STATOR

BEGINS TO CLOSE. WARM PORT

BEGINS TO OPEN. DISPLACER MOVING IN

WARM

COMPRESSION

SPACE

E

O

U

•? S ! ~o .

Figure 1 1 . Sequence of operations of Sunpower, Inc. proposed spin bearing demonstration model.

The turbine is positioned so as to intercept this maximum flow of gas at a brief interval during the outermost

point of its travel, thus imparting a spin to the displacer.

As the displacer moves back in again, it covers the stator, and once again uncovers the warm port so that normal flow through it can occur. Flow outward at its midpoint, and the piston has almost reached its outermost

travel. As the component motions continue, the flow through the warm port reverses and now flows through the port

toward the cold space, increasing in magnitude under the influence of the piston. The piston reaches its midpoint and

maximum inward velocity at about the time the displacer reaches its innermost point and reverses its direction to come out again and repeat the cycle.

During the entire cycle, the displacer is spinning, even though it receives a spin impulse only during the 45° of

its travel, and as it spins, it pumps gas through the brake after the manner of a squirrel cage fan. This brake is designed to limit maximum rotating speed to reasonable values fixed by bearing performance. The brake is only effective at high speed and has no retarding effect on the displacer during startup, when maximum turbine torque is applied to begin the spinning action.

4. Conclusion

The major effort in the Goddard program is to develop the magnetic bearing cooler. The self-actuated spin

bearing approach is still in the study phase; however, initial analysis shows that the approach is promising.

5. References

[ 1 ] Sherman, A., Gasser M., Goldowsky, M., Benson, G. and McCormick, J., "Progress on the development of a 3-5 year lifetime Stirhng cycle refrigerator for space," in Advances in Cryogenic Engineering, Volume 25.

114
[2] Sherman, A., "Cryogenic cooling for spacecraft sensors, instruments, and experiments," in Astronauts and

Aeronautics, November 1978.

[3] Gasser, at al., "An approach to long-life Vm cryogenic refrigeration for space applications," AppUcations of

Cryogenic Technology, Volume 4, 1971.

[4] Studer, P. and Gasser, M., "A bi-directional linear motor/generator with integral magnetic bearings for long

lifetime Stirling cycle refrigerators," Topical Conference on refrigeration for cryogenic sensors and elec- tronic systems, October 1980, at National Bureau of Standards.

115
THOR CRYOCOOLER

G. M. Benson, PhD and R. J. Vincent, PhD

ERG, Inc. Oakland, CA 94608

A resonant, free-piston Stirling-cycle cryocooler driven by

linear reciprocating synchronous electric motors and utilizing

gas bearings and gas springs has been designed, and selected

components tested. This hermetically-sealed, dynamically

balanced design is based on ERG's previous work and has been

engineered for long-life through use of: 1) non-contacting

linear gas bearings and narrow clearance reciprocating seals,

2) brushless reciprocating electric motors with hermeticallysealed coil and core, 3) stationary external first-stage

regenerator, and 4) solid-state controller. The design has

been efficiency-optimized by 1) transferring heat nearly

isothermal ly (to better approach Carnot efficiency), and

2) minimizing losses through use of advanced components of

resonant free-pistons (which eliminate piston drive losses).

The single-stage unit, designed to produce 10 watts at 65 K, may be expanded to three-stages, using a novel design, for

producing 4.2 K. Designs and test results of selected

components are described and compared with theoretical and

computer modeled results.

Key words: Cryocooler; gas bearings; refrigeration;

regenerator; resonant free piston; Stirling cycle.

116
A CRYOGENIC SYSTEM FOR THE SMALL INFRARED TELESCOPE FOR SPACELAB 2

E. W. Urban and L. Katz Marshall Space Flight Center

Huntsville, Alabama

J. B. Hendricks and G. R. Karr University of Alabama in Huntsville

Huntsville, Alabama

The cryogenic system for the small helium cooled infrared telescope

for Spacelab 2 is described. The system consists of a 250 £ storage

dewar containing superfluid helium at about 1.6K and a separate experiment cryostat. The liquid helium is confined within the dewar in zero-g

by means of a porous plug. The cold vapor leaving the plug surface is split into two, not necessarily equal, flows. One flow path is through

heat exchangers which cool the radiation shields of the storage dewar

and, therefore, control the liquid storage efficiency. The second flow

path is through a flexible joint to the independent set of heat exchangers

for cooling the infrared optical apparatus and radiation shields in the

cryostat. By means of commandable valves on the dewar and cryostat vent

lines, the fraction of total helium boil off used for cooling the dewar

radiation shields can be controlled, and the cooling available at the

cryostat can be adjusted from zero flow rate to nearly 50 mg/s of helium

gas at 2K. A number of unusual cryogenic problems which have been solved

in the development of this complex system are discussed.

Key words: Cryogenics, infrared telescope, Spacelab, superfluid helium.

The Infrared Telescope (IRT) Experiment will first be flown on .the Spacelab 2 mission

in late 1983. It is a joint project of the Smithsonian Astrophysical Observatory (SAO),

the University of Arizona (UA), and the NASA Marshall Space Flight Center (MSEC). Our

responsibility at MSEC is the development of the cryogenic and mechanical systems of the IRT. The University of Alabama in Huntsville (UAH) is collaborating on the cryogenic design,

fabrication of special cryogenic apparatus, and cryogenic performance testing. UA and SAO

are responsible for the infrared instrument and data analysis. The NASA Ames Research

Center is collaborating in the analysis of contamination and zodiacal light data. This

paper describes the cryogenic apparatus, the status of assembly and testing, and a number

of special cryogenic problems which we have faced during development.

The IRT experiment employs a very unusual cryogenic design which results from both the

requirements of the infrared science and a desire to utilize commercial apparatus and

fabrication technology wherever possible. Two views of a scale model of the IRT are shown

in Figure 1, and a schematic of the cryogenic system is shown in Figure 2. A description

of the mounting and operation of the 3.4m tall experiment in the Spacelab 2 payload was previously given [1] ,[2]. The two major cryogenic components, the dewar subsystem and the

cryostat subsystem, are described below.

The dewar subsystem consists of two parts, a 250 liter liquid helium storage dewar and

a transfer assembly (TA), which form a closely integrated thermal syster:,. The result is essentially a single complex superfluid helium dewar with the liquid containment vessel at one end of a neck structure and a number of cold flow control components at the other; all are surrounded by a common thermal protection system, discussed below.

1. Introduction

2. Cryogenic System

2.1 Dewar Subsystem

117
Figure 1. Two views of a scale model of the infrared telescope.

2.1.1 Dewar

In a conventional storage dewar of the type used, to ship helium on the highways the

liquid vessel is surrounded by a system of nested vapor cooled shields (VCS), separated by

superinsulation and mounted within a high vacuum space. These shields serve to intercept

heat being radiated and conducted in toward the cold liquid from the outer shell. The

intercepted heat is then conducted up each shield to the dewar neck where it warms the venting helium gas. The dewar neck is a heat exchanger as well as the support for the liquid

vessel. The source of the cold gas venting up the neck is the slow evaporation of the

stored liquid by the unavoidable parasitic heat conducted down the neck tube and radiated

in from the coldest VCS. A typical unmodified conventional 250 liter dewar achieves a steady state boiloff rate of 1% per day or 2.5 liters of liquid per day corresponding to a heat input to the liquid of 75 mW.

In order to utilize the basic features of a commerical storage dewar for the IRT,

several modifications have been required as shown in Figure 3. To retain the stored LHe

in the zero gravity of space, it is necessary to close off the dewar neck at the liquid

vessel entrance. Liquid is filled and withdrawn through tubes which pass down the neck

from the TA, and valving and flow control is accomplished by the cold components in the TA. These components are also insulated from the outside world by a VCS system, which is tied

to the dewar shields within the neck tube. The TA and the interior of the dewar neck

share a common vacuum independent of the dewar guard vacuum. Heat intercepted by the dewar

shields is conducted up to the heat exchanger in the TA.

118
KEY

B BURST DISC

F FLOW METER

H HEATER

L LEVEL METER

P PRESSURE GAUGE

QCM QUARTZ CRYSTAL MICROBALANCE

RV RELIEF VALVE

SS SUN SENSOR

T TEMPERATURE SENSOR V VALVE

V19

<V15

V16(RV3)

VENT

INTERFACE

(RV4)

liiiiii)iiiiiin\iiiiiiiii}iii/ii//iiiiiiiiii//}lllllMlllllJ

'PEWAR

VACUUM MAINTENANCE

ASSEMBLY — (ACTIVATED BY

UMBILICAL CONTROL BOX

(115V AO)

•ACTIVATED BY GSE DEWAR FILL CONTROL BOX

Figure 2. Schematic diagram of the infrared telescope system.

119
V20

DEWAR

BYPASS

VALVE

V17—

VACUUM PUMP-OUT PORT

WITH INTEGRAL RELIEF VALVE

V4-(RV6)

VALVE ACTUATOR COVER

GAS LINE HELIX CRYOSTAT LINEFERROFLUIDICSEAL

BELLOWS

POROUS PLUG

LOW CONDUCTIVITY

NECK TUBE (TYP)

RETRACTABLE VALVE

OPERATOR (TYP)

• ELECTRICAL CONNECTOR

VENT LINE RELIEF VALVE

V16(RV3)

VACUUM PUMP-OUT PORT"

WITH INTEGRAL RELIEF VALVE

V8(RV-5)

MULTILAYER

INSULATION

(TYP)

SUPPORT TUBELIOUID WITHDRAWAL/VENT

TUBE

LIQUID LEVEL

PROBE

LIQUID HELIUM

(1.6 K,6T0RR)

: TRANSFER ASSEMBLY

MULTILAYER INSULATION

VAPOR COOLED SHIELDS (TYP)

GHE VENT BAYONET

AND VALVE V13

.GHE VENT (TO FLOW

CONTROL VALVES)

FILL LINE RELIEF VALVE

V15 (RV4)

MOUNTING RING

-OUTER SHELL

SUPPORT

STRAP (3 @ 120")

T THERMOMETER

H HEATER

Figure 3. Schematic diagram of the IRT dewar subsystem.

120
When the experiment is launched in the Space Shuttle the dewar subsystem will be

horizontal, and large accelerations will be applied transverse to the dewar axis. To aid

in supporting the mass of the 250 liters of helium, the liquid vessel and the dewar VCS

against these loads, three f iberglass/epoxy straps attach the bottom of the liquid vessel

radially to the outer shell. Our dewar was fabricated by Cryogenic Associates, Inc. It has been vacuum leak tested, but cryogenic testing must await the assembly of the TA onto

the dewar.

2.1.2 Transfer Assembly

The TA, designed and fabricated by UAH, is currently being assembled onto the dewar

at MSFC. Within the outer case of the TA is a set of three nested heat exchanger tubs,

shown in Figure 4, each with a gas tube welded to its upper circumference. Each tub will

be wrapped with MLI and bolted to a dewar shield extending up from the neck. As described

below, a portion of the dewar boiloff gas leaving the porous plug, is circulated in turn through the three heat exchangers and then to external valves.

Within the inner tube is a demountable plumbing module which is supported by the dewar

neck. The module has five tubing connections, each utilizing a Varian mini conflat ultra

high vacuum flange. These connections are to the dewar fill and vent tubes, the external

fill line, the inner heat exchanger tub^and the gas line to the cryostat. The module

components, shown in Figure 3, operate at the temperature of the liquid helium, nominally

1.6K. They are described below.

Liquid helium is introduced into the dewar subsystem through a standard bayonet and vacuum jacketed valve (V6), then through cold fill valve (V7) and the dewar fill tube. The

cold fill valve (V7) prevents liquid helium from flowing out the fill line to the warm outer

shell when the experiment is in space. A burst disk (Bl) permits controlled venting through

the fill line in the event of accidental overpressurization. In space, liquid flows into

the TA through the withdrawal line and is restrained by the porous plug. The principle and

operation of porous plugs, which are essentially liquid/vapor phase separators for superfluid helium are described elsewhere [3]. In the event of plug blockage and overpresurization a second bypass burst disk (B2) provides relief into the vent lines. When the

dewar is being initially filled or when normal helium (4.2K, 1 atmosphere vapor pressure) is being converted to superfluid (1.6K, .008 atmosphere pressure), large gas flow rates must be

vented. A bypass valve (V5) is opened in parallel with the porous plug to permit high flow

rates and avoid possible contamination of the plug pores. A second bypass valve (V17) is

opened when liquid topoff begins to prevent initial warm helium from the warm transfer line

from entering the liquid vessel.

The cold valves, V5, V7, and V17 are operated by vacuum-tight, retractable actuator

shafts having low thermal conductivity and low thermal mass. When retracted, the actuator

shafts will be held in place by a cover box which can be evacuated through valve (V20) to prevent air leakage through the shaft retraction seals during the vibrations of launch. To reduce wear on the valve seats and minimize heat input when the valve actuators are engaged,

the valves will be opened at the beginning of cryogenic operations and will not be closed

except when the dewar subsystem must be turned on its side (launch attitude) for testing or just before the actual launch. Minor leakage past the seats of V5 and VU can be tolerated,

but V7 must remain sealed superfluid tight even in space to avoid heat conduction down the

fill line from V6. To prevent the cold valves from vibrating open during launch we have

designed a spring loaded detent lock operated by the valve actuator.

121

The liquid helium evaporates from the porous plug at a temperature of about 1.6K and

enters a manifold from which it is divided into the dewar and cryostat vent flows discussed

above. The cryostat vent tube is wrapped into a helical coil which flexes with the motion

of the cryostat as the infrared instrument is scanned across the sky. The insulated cryostat

line then passes through a ferrofluidic seal which joins the TA to the cryostat, accommodates

the scanning motion, and maintains the high vacuum within the cryostat and TA before the

experiment is carried into the vacuum of space.

The dewar subsystem has been instrumented to permit accurate measurement and control of

its operation on the ground and on-orbit. A number of thermometers measure liquid and VCS

temperatures. Thermometers T5 and T6 on the porous plug indicate the plug performance and

a commandable heater gives us a means of plug flow control in addition to the external

valves (see below). A thermometer (Tl) and liquid level probe are inserted into the dewar

from the TA and can be replaced in case of malfunction. Several superfluid tight electrical

feedthrus are used to bring instrumentation leads from the liquid helium out into the vacuum space.

2.1.3 Dewar Subsystem Thermal Performance

The dewar subsystem will be tested by itself at MSFC while the infrared instrument is

being installed in the cryostat in Arizona. Initially the gas line to the cryostat will be

terminated within the TA and the dewar tested as a stand-alone subsystem. The dewar will

be filled with normal helium and allowed to stabilize; then the helium will be converted to superfluid at about 1.6K by pumping through the vent line. A long term performance test

will be made with the dewar upright (porous plug not wet by the liquid helium), tilted 90"

into the launch attitude (plug may or may not be wet), and tilted beyond 90° to insure

that the plug is wet and operating as it will in space where fluid flows easily to it.

Actual performance will be compared to predicted performance.

Initial estimates of the thermal balance in this complex dewar led to the nominal

temperatures indicated in Figures 2 and 3: 20K for the inner VCS, 60K for the middle VCS,

and 120K for the outer VCS. The final configuration of the system, particularly the neck

tube structure and the support straps, has led to a revision of the analytic predictions

for steady state operation [2]. In the case in which all boiloff gas is used to cool the

dewar VCS (none to the cryostat), the shields are predicted to operate at 30K, 126K, and

240K respectively. The corresponding predicted heat load on the liquid vessel is 152 mW

and the boiloff rate is 6.7 mg/s, or 4.0 liter/day of superfluid helium.

When the stand-alone tests are completed, a cryostat substitute will be installed

on the cryostat gas line of the TA. This test equipment will allow us to simulate the

division of gas flow between the dewar and cryostat shields and to make preliminary tests

of the flight situation.

2.2 Cryostat

The cryostat, shown schematically in Figure 5, will be a special modification of an open neck laboratory dewar. The essential special features include mounting flanges for

the two sections of the infrared telescope, an access port for infrared detector installation, a side extension on the cryostat rotation axis through which the cold gas from the TA

enters, and a gas heat exchanger and VCS system. A coimiandable vacuum cover insures that a high vacuum can be maintained within the cryostat and telescope before the experiment is in

space. The cold gas from the TA is first delivered at a temperature of about 2K to a cold

finger, designed to maintain a temperature of about 2.5K, to which the IR detector block

will be clamped, and theh to the heat shield system. The lower and upper telescope sections

are designed to operate at maximum temperatures of 8K and 60K, respectively. They will be

cooled by conduction to their mounting flanges. The two telescope mount flange rings, and

two additional rings which are at temperatures of approximately 120K and 200K are suspended

from the top of the cryostat on large diameter f iberglass-epoxy tubes. Each ring carries a VCS and each delivers its collected heat to the cold venting helium gas. The vent tube,

after leaving the cold finger, is fastened to the four rings in turn, exits the cryostat at ambient temperatures, then connects to external flow control valves. Additional design

information was given previously [1]. The cryostat was also built by Cryogenic Associates,

Inc. 123
GHE VENT

(TO FLOW CONTROL)

VACUUM PUMP-OUT PORT

WITH INTEGRAL RELIEF VALVE

UPPER TELESCOPE FLANGE AND THERMAL

RING^eOK)

UPPER TELESCOPE TUBE

MULTILAYER

INSULATION 9-

LOWER TELESCOPE FLANGE AND THERMAL

RING I8K)

ACCESS PORT

OUTER SHELL

LOWER TELESCOPE TUBE

WARM THERMAL

RING (~200KI

LOW CONDUCTIVITY SUPPORT TUBES

MOUNTING RING

VACUUM COVER SHROUO AND

LATCH MECHANISM MOT SHOWN

ELECTRICAL FEEDTHROUGH

INTERMEDIATE THERMAL

RING (~120K)

GAS INLET LINE (2K) FROM DEWAR SUBSYSTEM

COLO FINGER I2.5K)

MOUNTING RING

VAPOR COOLED SHIELDS

® THERMOMETER

Figure 5. Schematic diagram of the IRT cryostat subsystem.

124
The cryostat was tested cryogenically, using a transfer assembly substitute which

permitted the controlled flow of cold helium gas into the cryostat line from. an ordinary

helium dewar. The results of these tests, discussed elsewhere [2], indicate that a flow

rate of about 17 mg/s will be required, at an inlet temperature and pressure of 2K and 6 torr (.008 atm), respectively, in order to maintain the infrared optical components at their

design temperatures.

2.3 Cryogenic System Performance

After the cryostat and dewar subsystems are integrated together, a cryogenic system

test will be run to verify the combined operation of the system. We will be particularly

interested in confirming the predicted steady state and transient behavior as we vary the

division of vent gas flow between the two heat exchangers. We must develop an optimum

strategy for the experiment operations prelaunch and on orbit in order to conserve helium

and still permit successful infrared observations whenever required. The only source of

cold gas is vaporization of liquid helium from the dewar/porous plug by either parasitic

heat, or electrical heat from the heater on the porous plug. The parasitic heat can, of course, be controlled over a wide range by varying the amount of gas in the dewar vent line. For a given rate of heat input, total pumping rate, and division of flow^the liquid will come to some equilibrium state. Regardless of the division of flow, total pumping rate

must be adequate to keep the liquid in the superfluid phase.

When the 17 mg/s required for infrared observations is diverted to the cryostat, the

dewar flow is predicted [2] to be only 3 mg/s. In this situation the dewar shields will warm to temperatures of 90K, 205K, and 260K respectively. The resulting total system

flow rate of 20 mg/s or 12 liter/day could be maintained only for about 21 days, if the

dewar were initially full of superfluid. With the Spacelab 2 on-orbit mission duration of

9 days and an anticipated prelaunch period after dewar topoff of as much as 15 days, it

will not be possible, nor, in fact, necessary to fully cool the cryostat until after launch.

As shown earlier, dedicating all flow to the dewar results in lowest dewar shield temperatures, a warm cryostat, and a total superlfuid boiloff rate of 4 liters/day or a full

dewar lifetime of 62 days. Clearly an intermediate situation will be chosen with the

infrared optics held at a fairly low temperature until the experiment is in space.

Adjustment of dewar and cryostat vent flows will be possible during the mission by

operating a set of three commandable warm valves in each of the vent lines as shown in

Figure 2. The three parallel on/off valves in each vent are sized with orifices allowing

14%, 29%^and 57% of full flow in the vent line. Thus, particular combinations of 0, 14, 29,

43, 57, 71, 86, or 100% of full flow can be selected for each vent line with the setting of

three valves. The system, and particularly the telescope and liquid helium temperatures

will respond rather slowly to changes in these valve settings. We anticipate the occasional need to rapidly deliver additional cooling to the cryostat as, for example, when

closure of the warm cover dumps a burst of heat onto the cold optics. This rapid

delivery of extra cooling can be achieved by activating a small heater on the vapor side of

the porous plug. The well-known superfluid helium "fountain-effect" [3] then draws liquid

through the plug.

3. Prelaunch Cryogenic Operations

Cryogenic servicing of the IRT experiment before launch presents a number of unusual

and challenging requirements. The Spacelab 2 payload will be installed into the Space

Shuttle Orbiter several weeks before launch. Perhaps 3 weeks before launch the IRT will be

purged, filled with normal helium, and allowed to stabilize. Since the supply dewars must

be located on workstands of limited weight carrying capacity at the sides of the Orbiter

bay, several small supply dewars will be required and the vacuum jacketed flexible fill

line will be 4.5 m long. Venting of the cold gas must be through long insulated hoses to prevent condensation and resulting dripping of water onto other experiments or orbiter

hardware. The helium will then be pumped down to and maintained in the superfluid phase by means of large, remote vacuum pumps. As late as possible in the ground processing flow,

nominally 15 days before launch, the dewar will be topped off with superfluid. Recent

superfluid topoffs performed at Ball Aerospace Division on the superfluid helium dewar for

the Infrared Astronomical Satellite (IRAS) have demonstrated that essentially 100% superfluid fill can be achieved.

125
The requirement that the liquid helium be held at a temperature and vapor pressure

below 2.17K and 38 torr (.05 atm), respectively, (the superfluid transition) means that

essentially continuous pumping must be maintained until launch. Pumping through long

umbilical vacuum lines to overboard pumps is not feasible for the Spacelab 2 mission, so we must provide our own on-board vacuum maintenance system, powered almost continuously

through an umbilical electrical line. This pump must operate satisfactorily when the

Orbiter is horizontal and also when it is in the vertical launch attitude. Through this

period the equipment is inaccessible. We are using a commercial, direct drive laboratory

pump and valves for this purpose. The pump was found to operate satisfactorily. over an angular range of about -50° to 45°, measured about the shaft axis from normal upright

laboratory attitude. The pump is therefore mounted tilted -50° on a bracket; when the

Orbiter is raised to the vertical, the pump is at 40° and still operates.

There will be a few unavoidable power outages for the vacuum maintenance assembly as the Orbiter is moved, raised to the vertical, mated to the External Tank and Solid Rocket

Boostersjand rolled to the launch pad. When the pumping is thus diasbled, the helium will warm; it must not warm above the superfluid transition. We have estimated that one fourhour power outage and several shorter ones can be tolerated, if adequately separated to give the system time to recool . Pumping will be continued until a few minutes before

launch. During ascent some heating of the liquid is expected, mainly due to the absence of

vent flow. Recent vibration tests on the cold IRAS dewar indicate that very little

frictional heating of the superfluid helium occurs. When the Orbiter reaches an altitude

at which the atmospheric pressure is below five torr, a barostatic switch will open a valve

between the experiment vent tube and the overboard vent. This will permit pumping on the

superfluid to resume several hours before the Spacelab is activated.

We will monitor and control the cryogenics during the mission from the ground in the

Spacelab Payload Operations Control Center (POCC). Emergency monitoring and control will

also be possible by the on-board Payload Specialists (scientists). Shortly before landing

the experiment will be stowed, the vacuum cover sealed, and the vent valves closed. The

relief valves on the vent lines will permit the experiment to warm up safely and untended.

4. Conclusion

The cryogenic system for the Spacelab 2 Infrared Telescope Experiment has been designed

as a low-cost approach to the storage of liquid helium during the preparation of a Spacelab

payload for launch and the supply of adequate cooling to a separate Infrared Telescope for

the duration of a typical Spacelab mission. The result is a system which will satisfy the

scientific requirements of the IRT and also be capable of furnishing cooling to a variety

of follow-on Spacelab experiments. It is expected that as the Spacelab and Space Transportation System flight program matures, flight durations will increase and prelaunch

preparation times will decrease. The superfluid helium system described herein should be

capable of satisfying experiment helium cooled refrigeration requirements for the next

several years.

5. References

[1] Urban, E. W., Katz, L., Karr, G. R., and Hendricks, J. B., Spacelab 2 infrared

telescope cryogenic systems. Society of Photo Optical Instrumentation Engineers,

Volume 183 - Space Optics, 1979.

[2] Karr, G. R. et al . , Cryogenic subsystem performance of the infrared telescope for

Spacelab 2, Proceedings Eighth International Cryogenic Engineering Conference,

ICEC-8, Genoa, 1980.

[3] Karr, G. R. and Urban, E. W., Superfluid plug as a control device for helium coolant.

Cryogenics, 266-270 (1989).

126
Requirements for and Status of a 4.2 K Adsorption Refrigerator Using Zeolites

William H. Hartwig

Department of Electrical Engineering

The University of Texas at Austin

Austin, Texas 78712

Previous work by the author and his associates has demonstrated

the feasibility of 77 K adsorption refrigerators using one N^O stage

which provides the adsorption temperature (185 K) for N^. Tne con- cept of adding an Hp stage and He stage has been examined theoretically

and experimentally using zeolites and/or other molecular adsorbers.

The results are described which shows the requirements for each stage

and suggests that a reliable low-power 4.2 K non-mechanical refrigerator

can be built with a considerable development effort and some additional

research.

Key words: Adsorption; compressor; cryogenics; molecular adsorption;

refrigeration; zeolites.

1. Molecular Adsorption/Desorption Compressors for Cryogenic Gases

It is well known [1] the clay minerals called Zeolites adsorb large volumes

o£ such gases as N2O, CO2 and NH3 at room temperature and pressure. For example,

the synthetic sodium aluminum silicate zeolite NaY will adsorb 21% by weight of

N2O at room temperature and pressure. The family of adsorption isotherms for

N2O and the zeolite NaY is shown in Figure 1 as an exanple. The adsorption pro- cess is exothermic, generating a heat of adsorption, H^, which is less than the

heat of vaporization of the normal liquid. The adsorbed gas density is, however,

approximately that of the liquified gas at its normal boiling point. For that

reason, addition of heat will release a large mass of adsorbed gas in a small

volume, creating a high pressure.

Cryogenic gases with lower boiling points can also be adsorbed with a useful mass fraction if the zeolite temperature is suppressed to about 2-3 times

their normal boiling points in absolute degrees. The exact pressure ratios and

desorption temperatures for nitrogen, hydrogen, and helium depend upon the choice

of zeolite. The practical problem is to find adsorbing media for helium at

The Molecular Adsorption Refrigeration System (MARS) is being developed to take advantage of several desirable features; no shaft seals or lubricant contamination, low- tolerance fabrication techniques, no mechanical vibration or acoustic noise background, long mean time to repair or failure, low cost materials,

and the ability to use waste heat.

a. Advantages of Adsorption/Desorption Gas Compressors

127
128
The disadvantages include a lower CO. P. than mechanical compressors, but

this is partially offset by the large compression ratio. To operate with lower

boiling point gases than N2O the adsorption temperature must be suppressed below

room temperature, but some of the heat can be cycled reversibly with proper design.

b. Theory of Gas Adsorption on Zeolites

The crystal structure of zeolite is that of an open lattice of Al, Si, and

0 atoms with dominant cations that may be Na, K, Cs,or other metals. The open

structure is a three dimensional periodic array of voids, typically 13A in diameter, with 5A openings between adjacent voids. Normally each of these voids fill

with 10-20 H2O molecules. When heated in a vacuum the water is desorbed, and

may then be cooled and filled with polar molecules of other gases . The adsorption of gas molecules on the interior surfaces of the voids is an ionic interaction with a characteristic potential energy for each different gas molecule at

each different ion site. The functional relationship of amount of adsorption

versus temperature and pressure for various' zeolite-gas pairs is extremely conplex,

and as such emperical methods have been commonly used to generate adsorption data.

Recent work by Woltman and Hartwig [5,6] has led to a theory of gas

adsorption derived from first principles. A lattice model of solutions is the

basis of the theory, and it treats adsorbed gas molecules and the vacant sites

where there are no adsorbed molecules as a two-component solution. The theory

is capable of accounting for the detailed physical behavior over a wide range of

pressures, and thereby establishes a reliable formalism to predict behavior and

design apparatus.

c. Closed Cycle MARS Compressor Configurations

While many configurations are possible, a typical one consists of four

cylindrical pressure vessels filled with zeolite. Each compressor is able to heat and cool the zeolite with a finned heat exchanger, using electric heat and

circulating liquid coolants. Each compressor successively adsorbs gas, pressurizes with addition of heat,desorbs gas, and then cools to adsorption pressure.

Figure 2 is a cycle diagram, where the vertical scale is heat power-, positive if

added and negative if removed.

In cryogenic refrigeration service the desorbed high pressure gas is delivered

to a Joule-Thomson orifice or expansion engine, and the low pressure return gas

is adsorbed. The adsorption cycle is illustrated in Figure 3. Gas is adsorbed

at 1 Atm. pressure up to point a, where x/m is 0.15 grams of gas per gram of

zeolite. Upon addition of heat, gas is desorbed and pressure rises to 21 Atm. at

b. Gas is delivered at 21 Atm until x/m reaches c. The cooling interval removes Heat and the pressure drops to 1 Atm. at d.

2. Compressor Design Theory

For a given mass flow, nu, the compressor delivers a fraction, Ax/m, of the

adsorbed gas from grams of^ zeolite in the desorption time, t^j. Hence

M (Ax/m)

^ ^d

To pressurize and desorb the gas, a fixed energy must be added to the compressor

per gram of zeolite, or

= K^CpAT (2)

129
B Adsorb;

I B

Pressure! Desorb

'I I B Cool

B Adsorb' Pressure

B Desorb E Cool

P -ic Adsorb!

I Pressure I Desorb C Cool [7r~, J , I Pressure

1 C Adsorbi

c Desorb

D

Desorb n Cool f D Adsorb

D

Pressure

D

Desorb

2t

Fig. 2 Heat Cycle diagram for MARS compressors. Each one,

A-D, is always in one of four states; (1) adsorbing

gas at low pressure, (2) pressurizing by addition

of heat with no flow of gas, (3) desorbing gas under

pressure, and (4) cooling the zeolite bed to reach

adsorption pressure.

where K]^ is a constant, Cp is the specific heat and AT is the zeolite temperature

rise. The average power input during the pressurizing and desorbing intervals is

q ,M K^M C AT ^ t +t , t +U ^ ' p d P d

Eliminating M2 between (1) and (3) , the mass flow of gas is proportional to the

input power, or

K^C AT

where K2 includes and tj/ (tp+t^j)

,

130
snm-j-rrr-rT-' ' » *

| M M | rm J

! r r f

f TT ( i 1

1 i i » | i i H-

(rtiass of N^O/mass of zeolite)

131
Desorption pressure, P^, is the dominant gas parameter that links gas flow

to cooling power. By referring to the T-s or H-P diagrams of the cryogenic

gases in their superheated regions, an emperical relation can be derived.

P,M A(x/m)

P> = AQ = (5) d g • o t^ ^

-

where A is a constant over a wide range of pressure. Equation (5) can be solved

for Qq,

P,M A(x/m)

Qo = At, t6)

d

The Coefficient of Performance is the ratio of cooling power developed by the

stage to the heat removal rate for adsorption, which means we divide Equation

(6) by Equation (4).

COP =

(7)

Good design practice will increase the constant, K, by reducing the relative

mass of non- adsorbing materials, such as the compressor shell, using isentropic

expansion instead of J-T orifices, and taking advantage of low Cp materials at low temperatures. As with conventional refrigerators, the COP of a multiplestage MARS is the product of the separate COP's.

MARS compressors would be optimized in part, by selecting the lowest possible

adsorption temperature to maximize the total adsorbed mass of gas; which in turn,

insures the highest desorption pressure for a given zeolite. This choice is

limited in practice to the boiling points of natural gases, such as N2O, N2,and

H2, but more favorable combinations may include Ne, argon, CH4, and CO2.

The optimum zeolite for each gas is the one that maximizes the combined

COP, not necessarily the individual COP's. At the time of this report no

definitive choices have been made between various synthetic and natural

zeolites, and also charcoal.

Table 1 summarizes typical values found so far, for the parameters in the

equations above, all using a J-T expansion orifice. The values selected for

the various constants are realistic approximations in the present state of the

art. The optimum energy budget for any design maximizes Pd(Ax/m)/AT, and it

is noteworthy to point out neither zeolite mass nor the cycle time appear in

Equation (7). A critical parameter, concealed in K, is the effect of free

volume in limiting the maximum value of desorption pressure.

Desorption pressure can be thought of as a function of gas temperature, free

volume (exclusive of the zeolite and adsorbed gas volumes) , and the mass of

desorbed gas. The adsorption isotherms show the adsorbed mass is almost

independent of pressure, so if more gas is "vaporized" by changing the zeolite

temperature, the pressure will increase. To the extent this is true, MARS

desorption pressure will be enhanced by limiting the volme of the external

tubing and valves, and reducing the interstitial volume between zeolite grains.

132
TABLE 1

Vjcto Adsorb

.

Desorb

.

Tf^TTTn OK

P

\^ci LJUJ

K"A. AXm

AT

(OK') rnp

Ax/m

AT

1 NzO 300 575 21 23.8 0.10 275 0 4 050 3.6x10" 4

2 N2 200 475 30 0.62 0.10 70^ 0 0067 3.6x10" 4

3 H2 77 200 25 1.3 0.009 1232 0 0018 7.3x10" 5

4 He 20 77 25 0.42 0.028 573 0 0052 4.9x10" 4

1. 70° is the temp, reduction load on the first stage. Reduction from 4750K to 270°K is by thermoelectric- cooled circulating fluid.

2. AT for H2 compressor may be less for some zeolites.

3. AT for He compressor can be less since is probably optimum at about 6OOK

and the zeolite need not be taken to 77°K.

4. COP of a working stage is experimental data. Energy for desorption was added by joule heat and dissipated to the ambient.

3. Conclusions

The data in Table 1 is not indicative of optimum performance because it was processed with equations for isenthalpic expansions. Some errors are possible,

although the data is consistent with published values[7] for common zeolites. It

is reasonable, we believe at this writing, that optimum COP's with the zeolites

tested will be about twice those reported for N2O, N2 and He. The very low

values for H2 are not consistent with the others, and other zeolites will increase

them. As it stands now, a 4-stage MARS could have an overall COP of about

10"^. This would produce 10 mw of cooling at 4.2K with about 10 KW input power.

a. Directions for Future Research

The Woltman-Hartwig model for adsorption and an experimental program for

verification is now appropriate to match all the cryogenic gases with particular

adsorbers. This will produce a better understanding of the physics of adsorption/

desorption systems, and permit the MARS compressors to achieve the maximum values

of A(x/m)/AT by improved design and temperature control.

b. Directions for Future Development

Many avenues for innovation and design remain for increasing the COP of

multistage MARS. The H^ stage has the lowest individual COP, and it can be

increased by adsorbing at higher pressure, finding more effective zeolites,

or substituting other technology. A LaNi5 compressor which utilizes the

reversible hydride formation is one possibility now being studied elsewhere [8]. It would operate with a room temperature heat sink but needs a very small 77°K

stage to get the H2 gas below its inversion point. This could be supplied

with a two-stage MARS. If metal hydride compressors prove to be feasible,

the marriage of the two new technologies will be straightforward.

Many areas for compressor design improvements must be pursued to optimize the

cycle and give proper attention to reliability and maintainability. Work by the

author and his associates [9] has led to optimizing the heat exchanger geometry

for MARS compressors. Other problems involve the logic of valve networks for

routing gases throughout the system, and design of the valves themselves. The

best design approach appears to be rotary ganged valves that synchronize the

cycle in each stage.

133
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134
The use of simple expansion engines to improve the COP of each stage is an obvious choice. Miniature cylindrical or spherical rotors with gas bearings

are possible configurations, but the development effort will be significant.

Figure 4 is a schematic diagram of two He cr/ocoolers which compares JT and

rotating expansion.

A most in^ortant problem, for a multistage MARS design that is otherwise

feasible, is heat management. Advantage can be taken of heat leaks and waste

heat in hotter stages to provide desorption heat in colder stages. If the

optimum gas/zeolite combination is found for a particular stage the final

desorption temperature can be controlled to achieve the optimum P(j(Ax/m)/AT

and COP.

4 . Acknowledgements

The work was supported by NASA-LBJ Space Center, Houston, and by Rockwell

International, Anaheim, by a contract with the Navy. The author acknowledges

their contributions to the work.

5 . References

[1] D.W. Breck, Zeolite Molecular Sieves , John Wiley, New York, (1974)

[2] Hartwig, W.H. , Steinfink, H. , Masson, J. P. and Woltman, A.W. , Proc. 1976 Re- gion V IEEE Conf. , IEEE Cat. 76CH1068-6Reg5, 1976, p. 80.

[3] Hartwig, W.H. , "Adv. in Cryogenic Engineering," Ed. by K.D. Timmerhouse,

Plenum Publishing Co., N.Y. , 1978, 23, 435-447.

[4] Hartwig, W.H., Woltman, A.W., and Masson, J. P., Proc. 1978 Int'l. Cryogenic

Eng'g. Conf . , IPC Science and Technology Press Guildford, Surrey GU2 5AW,

England.

[5] Woltman, A.W. , Ph.D. Dissertation, The University of Texas at Austin, Austin,

Texas, March 1978.

[6] A.W. Woltman and W.H. Hartwig, "The Solution Theory Modeling of Gas Adsorption

on Zeolites," ACS Symposivm Series 135, pp. 3-25, W.H. Flank, Ed., (1980)

[7] A.J. Kidnay and M.J. Hiza, "High Pressure Adsorption Isothenns of Neon,

Hydrogen and Helium at 760K," Adv. in Cryogenic Engineering , Paper K-5, pp.

730-740, (1978), K.D. Timmerhaus, Ed., Plenum, New York.

[8] William Steyert, Los Alamos Scientific Labs., (personal communication of a paper submitted for publication)

.

[9] R.M. Flores and W.H. Hartwig, "Optimization of Heat Exchangers in a .Molecular

Adsorption Refrigeration System," Rpt. No. TSR-3, Electrical Engineering

Dept., The University of Texas at Austin, Austin, Texas 78712, (May 1980).

135
COOLING OF SQUID DEVICES

BY MEANS OF LIQUID TRANSFER TECHNIQUES

FOR REDUCED HELIUM CONSUMPTION

AND ENHANCED TEMPERATURE STABILITY

El don A. Byrd

Naval Surface Weapons Center

White Oak, Silver Spring, Maryland 10901

R. G. Hansen

R. G. Hansen & Associates

Santa Barbara, California 93101

In an effort to establish a state-of-the-art SQUID device for

both laboratory and field applications, a developmental effort has been

undertaken by Naval Surface Weapons Center to utilize a liquid helium

transfer system to perform cooling of SQUID devices in the temperature

region of 4.5 to 16 Kelvin. Various techniques have been employed to enhance the temperature stability of the inherent system as well as to provide increased system portability and flexibility of application.

The specific refrigeration system used in this program has been

further modified for expanded applications in the low temperature testing of infrared sensors, low noise amplifiers, and other devices.

This particular developmental effort has resulted in a system

which allows the user to function in either a horizontal or vertical

orientation. Operation had previously been limited to single axis

orientation as a result of liquid cryogenic systems.

This paper presents a concept of utilizing state-of-the-art

cryogenic systems to interface with diverse technical needs in the area of electro-optics, SQUID devices, and supporting electronic

accessories.

Key words: Cryogenics, helium, infrared, SQUID, temperature sensors

This presentation, although originally intended to be a report on actual performance

results of a functional system, is respectfully submitted as an interim report on "results

to date" of a recently delivered system designed for both laboratory and field operation

of SQUID devices.

In the initial definition of the cryogenic support requirements, considerations such

as package size, system efficiency, cost of operation, and portability were defined as primary objectives. The specific function for the system was to perform the necessary

refrigeration to operate SQUID magnetometers in both laboratory and field experiments.

Secondary application such as infrared detector cooling, sample cooling for cryospectroscopy, and the testing of low noise amplification systems were also envisioned for

future use with the basic cryogenic system.

136
A refrigeration capacity of 0.250 watts at 4.2 Kelvin was the required refrigeration

capacity with a goal of 2 watts at 4.2 Kelvin. Temperature indication and control over the

entire range of 4.2 Kelvin to 300 Kelvin was necessary with desired stability of - 0.001

Kelvin over the specific region of 4.2 to 16 Kelvin.

Refrigeration techniques such as closed cycle refrigeration systems, traditional

dewars, and liquid transfer techniques were reviewed.

With all desired parameters being considered in view of available systems, the Air

Products and Chemicals' Model LT-3-110 Heli-Tran liquid transfer refrigeration system was selected. Of particular importance in this selection were the following rather unique

features of the Heli-Tran as it related to our needs.

1. The Heli-Tran's compact size and ability to operate in either a horizontal or vertical orientation was beneficial.

2. The "off-the-shelf" system was easily adaptable to our specific

needs without costly alterations.

3. Rapid cooldown of samples could be effected within 45 to 60 minutes

with cryogenic consumption of k one liquid liter of helium per

hour of operation.

4. The "0" ring interface of the Heli-Tran cold head was readily

adaptable to various vacuum shroud configurations, thus allowing

versatility of applications within varying technologies.

5. Temperature control could be effected with rapid response time

due to the "fine tune" control ability of the liquid helium

transfer and the cold head tip heater circuit.

The commercially available Heli-Tran LT-S-llO liquid helium transfer system is shown in

Figure 1 schematically. The functional system is shown in Figure 2.

In briefly summarizing the operation of the Heli-Tran, we provide the following:

A specially designed flexible transfer line is inserted into any standard

helium storage container. An insertion tube on the dewar end of the line

contains a fitting for pressuring the dewar from 3 to 5 psi and incorporates

a safety relief valve. The line itself consists of two very small flexible

tubes, one inside the other, and a separate return tube which acts as a heat

shield. Operation of the line is as follows: liquid is transferred to the

specimen through the innermost passage of the two concentric tubes at the

pressure of the dewar. To ensure 100% liquid and extremely steady flow in

the center tube, a second stream of helium is transferred at a slightly lower

pressure through the annular space between the concentric tubes. This lower

pressure (and therefore lower temperature) liquid is a result of flashing the

liquid to a slightly lower pressure by means of a capillary tube through which

the liquid enters the line from the dewar. Boiling and flow oscillations thus occur in the annular flow passage while the liquid in the center tube remains

sub-cooled. The liquid to the specimen is expanded at the cold end through

an adjustable needle orifice. This regulates the flow rate and therefore

refrigeration to accommodate varying heat loads. The exhaust vapor from the

specimen cools the radiation shield surrounding the specimen and is then heated

as it vents from the cold end to avoid frosting. The pressure of the exhaust

helium from the specimen may be lowered by connecting a vacuum pump permitting

temperatures as low as 2 Kelvin to be achieved.

The shield liquid in the annular space is vaporized by the heat leak into the

line. This consists primarily of conduction through the support system and

radiation. Radiation is minimized by use of superinsulation around the con- centric tubes. The cold gas from the annular flow passage is returned

137
through a tube parallel to the insulated concentric tubes. This further

decreases radiation and support conduction losses by using the sensible heat

of the shield flow. This gas is vented at the dewar end of the transfer line.

Temperature is monitored and regulated through temperature sensors and a heater mounted on the cold tip. The top of the conductively cooled unit

has. a l/^"-28 UNF thread in a high conductivity copper tip at the cold end

which permits a variety of specimen holders to be easily interchanged. [1]

TRANSFER

LITJE \

/ ' //

DEWAR

ADAPTER

MICROMETER VALVE

ADJUSTMENT

HELI-TRA1L_^

COLD END

1 . HELIUM DEWAR

2. HELIUM CYLINDER

3. PRESSURE REGULATOR

4. MERCURY MANOMETER, 0 to 20" Hg OR PRESSURE GAUGE

ON HELIUM DEWAR

5. VACUUM SHROUD

6. VACUUM SHROUD PUMP

7. VACUUM PUMP FOR OPERATION

BELOW 4.2°K

8. ACCESSORY FLOW CONTROL PANEL

9. ACCESSORY TEMPERATURE CONTROLLER,

MANUAL OR AUTOMATIC

FIGURE 1 Typical Test Set Up [2]

In the interface* of our particular SQUID device to the standard Heli-Tran, it was necessary to design a vacuum shroud approximately 46" in length so as to accommodate the

required electrical feedthrough. This was accomplished by means of a two-piece telescoping

vacuum shroud. v

The heat loss or load for the electrical feedthrough was calculated to be 0.230 watts. This heat load was received on the incoming cable by means of a thermal attachment of the

cable to the exhaust gas radiation station thus reducing the base load on the refrigeration

system and effectively utilizing the refrigeration provided by the exhausting helium gas.

To provide for the temperature indication and control of the SQUID, a gallium arsenide

diode thermometer was initially selected. Although not offering the most favorable

138
139
sensitivity, this type of thermometer provided the most favorable non-magnetic interference

for the SQUID.

As a secondary means of temperature monitoring, an iron-doped gold vs. Chromel

thermocouple was also inserted in the cold head. At the present time we are installing

a calibrated silicon diode thermometer into the SQUID interface in an effort to ascertain

the ultimate stability of the system over the temperature region of 4.2 to 16 Kelvin.

To perform the required temperature display and control, the Scientific Instruments

Incorporated Model 3800 Indicator/Controller was used. This particular instrument is com- patible with either silicon diodes or gallium arsenide diodes and provides proportional

heater output specified by the programmed temperature set point.

The salient parameters of the controller as it applies to our primary application

(4.2 K to 16 K) are as follows:

1. Readout Resolution and Repeatibil ity (Kelvin)

± .002 K from IK to lOK

i .010 K from lOK to 20K

2. With Si-400 Silicon Diode the Readout Accuracy- in Kelvin is ± .005 K from IK to lOK

± .010 K from lOK to 20K

3. Control Stability at Fixed Set Point with Constant Load and Ambient

(in Kelvin units)

± .001 K i .04% of temperature for one (1) minute

* .010 K ± .04% of temperature for one (1) hour

4. Power Output

One (1) ampere at 30 volts

(maximum for 46 ohm 20 watt heater integrated into Heli-Tran

cold head.) [3]

Having had but one opportunity to functionally test the above system, we accomplished

a cool down from 300 Kelvin to 11 Kelvin within ninety (90) minutes after commencement of

cryogenic transfer. Unfortunately the desired temperature could not be achieved due to a later discovered vacuum leak in a braze joint on the vacuum shroud. Temperature stability,

however, over the temperature range of 11-17 Kelvin was recorded with a stability of

greater than * 0.01 Kelvin; this represented the GaAs readout accuracy.

At the present time the vacuum leak is being corrected and the additional silicon

diode is being installed on the interface between the refrigeration assembly and the SQUID

device.

References

[1] Meier, R. N. and KOURY, A. G., "Cryogenic Systems for Research at Helium Temperatures"

Air Products & Chemicals, Inc., Advanced Products Department, Allentown, PA 18105

[2] Operation Manual for Model LT-3-110 Liquid Transfer Heli-Tran Refrigerators , Air

Products & Chemicals, Inc., Advanced Products Department, Allentown, PA 18105

[3] Scientific Instruments Incorporated Data Sheet Number 301277 entitled "Model 3800

Temperature Indicator/Control ler"

.

140
A CONTAMINATION FREE COMPRESSOR FOR SMALL SCALE

STIRLING REFRIGERATORS

J.G. Daunt*, C. Heiden

Institut fiir Angewandte Physik der Justus-Liebig-Universitat Giessen

Heinrich-Buff-Ring 16, D-6300 Giessen, West Germany

Contamination of the working gas appears to be' a major problem encountered in the design of reliable small

scale Stirling cycle refrigerators for cooling cryoelectronic devices over long uninterrupted periods.

Seals for the reciprocating shafts of such a refrigerator may be a source for contamination due to lubrication and possible small leaks. The use of bellows provides the potentiality for a hermetically sealed refrigerator. As a first step towards such a system, a bellows compressor was designed. For increased lifetime,

mechanical load on the bellows is reduced significantly by keeping the differential pressure acting on the

membranes at a low level. This is achieved by placing

the bellows in the volume of a conventional piston

compressor. Such a compressor system is tried in con- nection with a three stage plastic displacer unit.

First results on the performance of the scheme are reported and discussed together with possible design

variations that may lead to a further reduction of ' mechanical vibrations and magnetic interference of

the cooling unit.

Key words: Bellows; compressor; contamination; continuous operation; cryocooler; low power refrigerator; Stirling refrigerator.

1 . Introduction

Low Power Cryocoolers such as developed by Zimmerman et al. [2],

[3] are very attractive for use with cryoelectronic components such as SQUIDs or other devices based on the Josephson effect. Long uninterrupted

working periods of such a cooler are desirable not only when needed in con- tinuous measurements but also from the point of view of the rather long cool

down times of the order of several hours of the refrigerator. It may for

example be more convenient to leave the cooling unit for a magnetocardiographic system using a SQUID running continuously than always turning it on many hours before actual use. Since one of the major problems encountered

with the design of a continuously running small scale cryocooler is contamination of the working gas, efforts towards a hermetically sealed system were undertaken using bellows as a means for sealing. This paper reports first

experiences made during this work.

141
2. Basic Low Power Cryocooler Design

Cooling

tubes "^(UJ

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Basic Stirling cryocooler

design

.

Our Stirling refrigerator is based

on a design similar to that of Zimmerman [ 1 ]

• The three stage displacer and surrounding cylinder are made of glass fibre enforced plastic

(Cadillac G10) with close fit (cf.

fig. 1). The lowest stage of the

displacer has a bore which is filled

with lead shot in order to increase

the low temperature heat capacity.

The displacer unit is connected to a small compressor which consists of a graphite piston moving with close

fit in a copper cylinder. The com- pressed gas is cooled by a package

of air cooled copper tubes located

at the exit of the compressor. The

piston shafts of displacer and com- pressor are sealed with double

0-rings. The buffer volume between

the 0-rings is filled with helium

gas of the same average pressure as

in the compressor. The whole system

is connected via needle valves to a helium supply cylinder.

Running this machine with 5 bar

average helium pressure and a speed

of 115rpm with a compression ratio

of about 3 : 1 resulted in an almost

linear initial cool down behavior . with a temperature drop 120K/h,

measured at the bottom of the lowest

displacer stage. The final temperature was 24k. a reduction in average

pressure by a factor of two then re- sulted in a further decrease of the

lowest temperature towards a. value

near 18K. This is somewhat higher

than the final temperature that was reached by Zimmerman and coworkers [l ] We expect to obtain a lower temperature after optimization of the

present displacer unit.

After an uninterrupted running

period of three days, a noticeable

reduction in performance however

took place resulting in a slow temperature increase of the coldest stage.

The origin of this was found in the

formation of deposits at the lowest

displacer stage causing increased

friction. A possible source may be

small leaks due to the 0-ring seals

and also due to some remaining porosity of the plastic material in the

displacer unit that lead to losses of

helium gas. They are compensated by

connecting the refrigerator at appro142
I priate locations with high flow impedance (needle valves) to the pressure cy- linder of the helium supply. This however requires helium gas of high purity. Otherwise the refrigerator will act as a cryopump on the contaminations in the

gas eventually leading to the observed deposits. The transport of air molecules through the 0-ring seals along the greased shafts also has to be con- sidered in this context.

3. Cryocooler Using a Bellows Compressor

Since contamination appears to be a fundamental problem with regard to

long uninterrupted working periods a straight forward solution to it should

be the construction of a hermetically sealed system or at least one that comes close to it. As a first approach we started to build a contamination

free compressor. A simple well known means to separate different media for

instance in pumps is the use of bellows. Using this principle the present

set was modified by inserting a bellows into the piston compressor as shown

in fig. 2. In addition to being rather simple, this design has the advantage

of keeping the mechanical load on

the bellows at a low level which

should be of importance for a long

lifetime of the bellows. Except for

a difference due to the mechanical

stiffness of the membranes, the pressure inside and outside the bellows

is identical as long as it can breathe

.

The experience with a stainless

steel bellows used in the first ex- periment however was disappointing.

Driven with a compression ratio of

about 2 its lifetime was about 10^

strokes, terminated by fatigue fracture along one of the welded seems.

8 A lifetime of 10 strokes on the other hand would be quite appropriate. A lifetime exceeding this is

quoted for commercial bellows made

of teflon. Using such a bellows the

compressor has been operated with an average helium pressure of 5 bar and

without apparent degradation in performance now for a period totaling

more than 4 million strokes. Although

the compression ratio due to into Displacer unit creased dead volume in the compressor

was reduced to a value of 2.3 resulting in a somewhat slower cool down

Fig. 2: Sealing Bellows inside ^^^^ °^ 80K/h, the final tempePiston Compressor rature again was close to 18K. A

problem encountered during the cool

down period is the change of average

helium gas pressure inside the bellows, which has to be accounted for

by a corresponding reduction of the outside average gas pressure. So far

this was done manually, but should be effected automatically in a more elaborated design. Much longer testing periods of course are needed before anything definite can be stated about the long term performance of the com- pressor system, and there may be still other materials superior to teflon

for the bellows.

143
However, before any long period

testing is reasonable, the other

weak spot in the system has to be

eliminated which is the seal at the

displacer shaft. This can be done

using the same principle (cf. fig. 3)

To reduce again mechanical load

acting on the membranes the sealing

bellows is surrounded by a pressure

compensation cylinder connected to

the outside volume of the compressor

bellows. The time constants for filling the volumes of the displacer

unit and of the compensation cylinder should of course have about

the same value in order to avoid

larger dynamic pressure differences at the second, bellows

.

Fig. 3. Hermetically sealed Stirling

cryocooler

.

Refrigerator with Pneumatic Bellows Compressor

1 t

to from

Remote compressor

Instead of using a piston compressor to

provide the driving

pressure of the bellows , one could em- ploy a pneumatic

system as shown in

fig. 4. The compressor bellows is activated by periodically pressurizing and

depressurizing the

surrounding cylinder

via two valves that

are controlled by

the crankshaft of

the displacer unit.

This can be done

either mechanically

as shown or by magnet

valves. We tried a

Fig. 4. Stirling re- frigerator with

pneumatic bellows

compressor

144
setup with the latter valve type and found that such a system can be operated successfully with practically the same cooling performance as with bellows in the piston compressor. The inherent advantages of such a system are increased flexibility of design which allows for instance to place sources

of electromagnetic interference and of mechanical vibration at locations

remote from the cold end of the displacer unit.

Acknowledgements: Part of this work was made possible by a Senior US

Scientist award granted one of the authors (J.G.D.) by the Alexander von

Humboldt Foundation.

5. References

[1] Zimmerman, J.E. , Radebaugh, R. and Siegwarth, J.D. , Possible Cryocoolers

for SQUID Magnetometers, Superconducting Quantum Interference Devices

and Their Applications. Hahlbohm, H.D. and Liibbig, H., eds., Walter

de Gruyter, Berlin (1977) p. 287.

[2] Zimmerman, J.E. and Radebaugh, R. , Operation of a SQUID in a Very LowPower Cryocooler, Applications of Closed-Cycle Cryocoolers to Small

Superconducting Devices, Zimmerman, J.E. and Flynn, Th.M., eds.,

US Dept. of Commerce, NBS Special Publication 508, Washington DC

1978, p. 59

[3] Zimmerman, J.E. , Cryogenic Techniques for SQUIDs, Proc. of the second

intern. Conf. on Superconducting Quantum Devices, Berlin 1980,

in print

* Present address: Physics Department, Queen ' s University, Kingston,

Ontario, Canada K7L 3N6.

145
A BI-DIRECTIONAL LINEAR MOTOR/GENERATOR WITH INTEGRAL MAGNETIC BEARINGS

FOR LONG LIFETIME STIRLING CYCLE REFRIGERATORS

Philip A. Studer and Max G. Gasser

Space Technology Division Goddard Space Flight Center

Greenbelt, Maryland

Linear motors are well suited to drive Stirling Cycle machines since they reduce the machine dynamics to straight line motion. The first device is being constructed as a dynamic

balancer for the Stirling Cycle cooler, which is described in a companion paper at this conference. The motor concepts also applies to Stirling Cycle cryogenic coolers, and its advantages

in this application are presented. The permanent magnets, pole pieces and drive coils, which

form the stator, are stationary and external to the working fluid. The only moving part is the

iron piston, which is not in physical contact with anything except the working fluid. It is magnetically suspended and positioned to within 10 microinches by differential capacitive

sensors. This allows the clearance between the piston and cylinder to be small enough to form a clearance seal. The piston is also driven electromagnetically and its axial position sensed capacitively. Heat generated in the motor coils does not enter the working fluid. Gas

springs allow a controlled resonant drive to minimize power. The same magnetic structure is used for both the drive function and the magnetic suspension. The electromagnetic design

along with the sensing and control implementation are described in the paper. Preliminary

test data is presented. Magnetic bearing performance characteristics are discussed.

Key words: Clearance seal; gas compressor; linear motor/generator; magnetic bearings; Stirling

Cycle.

1 . Introduction

The hnear motor/generator is the heart of a dynamic balancer which is one element of a long-Ufe mechanical

refrigerator system being developed by the Goddard Space Flight Center. The need for cooHng capability in space

is growing with the use of larger sensor arrays and larger optics as well as future plans for cooled instruments and

superconducting devices. Long maintenance-free life is the prime goal since satisfactory short term missions «an be

accomplished with stored cryogens and satisfactory refrigeration systems are available where maintenance can be

easily provided. The goals of this development are directed toward technological advancement of the state-of-theart as well as providing an effective balancer. To illustrate, the motor concept was the result of an analysis of a basic limitation of hnear moving magnet motors and an effort to minimize the mass of the moving element. It establishes a basis for significant system improvements in future refrigeration systems. The major thrust of this overall development ii to ehminate all life limiting mechanical wearout phenomena by employing non-contacting

bearings and seals and reducing the number of moving elements to a minimum. The first step in this direction is the hnear drive motor which does away with the rotary to Hnear conversion mechanism. There is no crankshaft,

connecting rod, or wrist pin, each of which requires a set of bearings. Also to this end,magnetic bearings, a de- velopment pioneered by the GSFC in the early VO'sl^l for rotary devices, plays a major role. The close clearances imposed by the gas sealing requirements represent a new challenge in bearing design. Good thermodynamic per- formance requires additional analysis and tradeoffs to operate efficiently with clearance seals. The initial motor

analysis rejected some motor types based on highly destabiUzing side forces and focused on moving coil and moving magnet d.c. motors. Moving coil motors, which offer lower moving mass (at the expense of power) were dis- counted because of the need for flexing leads, a reliability hazard. A generalized expression for motor weight and

power was programmed (Figure 1) based on fundamental motor equations.

146
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EFFICIENCY

Figure 1. Typical Plot for Linear Motor Sizing

I

This plot is representative of the way power dissipation and the weight of the motor can be expected to vary

for a given set of parameters meeting a particular refrigerator requirement. This procedure established a minimum

weight for a given thrust and efficiency with currently available permanent magnet materials with the clear choice

of rare-earth magnets. The flux density of these magnets is one fourth the saturation value of magnetically soft

steels. These considerations led to the concept about to be described in which the piston is magnetically soft steel and both the permanent magnet and the drive coils are stationary.

2. Linear Motor Description

A permanent magnet biased electromagnet approach was selected. This allows the use of the highest magnetic

flux densities, on which the force generation is dependent. The piston, illustrated in Figure (2) is essentially a sohd iron bar with the maximum magnetic flux limited only by saturation of the central cross sectional area. This

determines the peak axial force which can be generated. In the final design the two end portions are composed of

thin ring laminations to reduce magnetic losses.

Figure (3) shows the bare stator structure which is symmetrical both axially and radially. This photograph

was taken prior to coil winding to show the structural design. A pair of poles sandwich the permanent magnets at each end and produce a high airgap flux in the major diameter at each end of the piston when it is inserted in the

bore. There is some magnetic axial centering force. The permanent magnet and pole assemblies are joined by four

bridging elements whose functions will be explained later. A rather thick wall non-magnetic cylinder is interposed

between the pole surfaces and the piston.

The magnetic circuit and typical flux paths are seen more clearly in Figure (4). On the right is a cross section of the stator assembly with the piston in place and the motor coil identified.

147

PERMANENT MAGNETS (8)

MOTOR COIL

MAGNETIC SUSPENSION LINEAR MOTOR

(AT MID STROKE)

PERMANENT MAGNET FLUX PATHS

CONTROL FLUX PATHS AND RESULTANT FORCE

Figure 4. Cryogenic Cooler Balancer at Mid Stroke

Note that none of the permanent magnet flux passes axially through the length of the piston when it is at mid-stroke. The flux from the permanent magnets is radially outward in both of the central gaps which are also part of the motor coil circuit. When the motor coil is energized the flux level on one side increases with a decrease on the other resulting in an axial force. A good motor should also be a good generator and this equivalent reciprocal action is evident in Figure (5).

Here the piston is shown displaced to the left which causes permanent magnet flux to pass down the length

of the piston following a path which encircles the motor drive coil. This change of flux occurring in the bore of

the coil generates a voltage proportional to the rate of change. The magnitude of this flux change from one end

of stroke to the other can be equal to twice the saturation flux level of the piston mid-section. And the midsection of the piston can nearly fill the bore of the motor coil. Therefore, a highly effective generator and motor

action occur. The efficiency of static force generation is the ratio of mechanical power generated to the input

power and the differences in the electrical input and mechanical output is dissipation which results in heating. The

drive coil resistance can be kept small because of the large volume available for the coil winding and, in contrast to rotary motors, there is no wasted energy in "end turns." All of the copper is effective at all times so that the electromagnetic efficiency can be very high. All of the stator structure in the final design is laminated to reduce magnetic losses. The permanent magnets are Samarium Cobalt, which has a coercive force ten times that of

^nicos' which makes it possible to drive large airgaps.

On the left of Figure (5) you see that each of the poles is actually four elements forming a cruciform shape

around the centered piston. The large magnetic gap of 0.25cm is shown to scale. The piston, when radially

centered, is at a point of unstable equilibrium. Huge side forces would be developed if the piston were free to move close to the poles in any direction.

The requirements of the long life refrigerator however, demand clearance seals and the piston is physically

limited to move less than ±0.003 cm within the cylinder wall. This means that a relatively thick wall pressure

cylinder can be interposed between the moving piston and the magnetic poles of the stator assembly removing all of the windings and magnets outside the pressurized working fluid. This is advantageous from several standpoints;

contamination, heat rejection, minimization of dead volume and the structural weight of the pressurized envelope.

149
MAGNETIC SUSPENSION LINEAR MOTOR

(END OF STROKE)

PERMANENT MAGNET FLUX PATHS

CONTROL FLUX PATHS

Figure 5. Cryogenic Cooler Balancer at End of Stroke

3. Magnetic Bearing Description

Referring back to Figure (4) left, note that the permanent magnet flux is always similarly directed radially

from each of the four surfaces making up the cruciform shape (in this case inward). The suspension control coils are mounted on the ring which connects the outer ends of the cruciform. When a pair of these coils are energized, the normally symmetric flux distribution is altered with an increase in two sectors and a decrease in the other two. A force is generated in the diagonal vector shown which is proportional to the product of the permanent magnet

flux and the control fluxl^l. This force is directly proportional to the control ampere turns and, just as in this motor, provides linear control at high efficiency. The efficiency is due to the permanent magnet bias flux level for which no power is required.

Active (servo controlled) magnetic suspension is required to obtain the precision necessary to center the close

fitting piston within the cylinder bore. Stiffnesses of the order of 1 X 10^ Newton per meter are required to overcome static and gas dynamic loads anticipated while maintaining zero contact (displacements less than 20

micrometers). Also required are means to sense the displacement of the piston relative to the cylinder wall.

4. Sensing and Control

We have selected differential capacitive sensing to measure and control the wall clearances. Capacitive sensing

is compatible with the large area short stroke requirements. The piston surface itself acts as the passive plate. Figure (6) illustrates the etched capacitor plates which are formed on the inner bore of the cylinder.

There are eight of the smaller plates encircHng each end of the piston. Four of these are electrically common

and are excited by an A.C. source in the mega Hertz range. The other four form two pairs of sensors whose coupling varies as a function of piston clearance. The inverse action in opposing gaps linearizes the output. A simple

demodulation circuitl^l developed by H. Machlanski of Mac Bar Mechanisms, Inc., has been found to be a very

effective signal source. Figure (7) shows schematically the four identical radial control channels and a diagramatically similar axial drive servo.
Figure 7. Momentum Balancing System Sensor Circuitry 151
The two sets of longer plates at each end receive the same excitation and provide an axial position output. For a resonant system only an output phase-shifted from the displacement signal is needed to maintain oscillating

motion, in which case the input power is very much smaller than the instantaneous power of the piston. When

the piston is doing work on the compressed gas, of course, drive power must be supplied to equal the work done

per cycle and the losses. The device currently being developed at GSFC is intended to be used as a dynamic

balancer for the Cryogenic Cooler which is the subject of another papert'') at this conference. The motor drive for the balancer will receive its input from an accelerometer with closed loop control to force the net vibration of

the cold finger to approach zero. It will be a resonant system with its piston moving 1 80 degrees out of phase

from the combined center of mass of the power piston and displacer. The control action will force the balancer

piston to assume the inverse of the velocity-displacement profile of the machine.

The selection of capacitive sensing avoids the need to penetrate the cylinder wall but requires a dielectric cyhnder. We have selected pyrex glass as the gas barrier with vapor deposited capacitor plates on its outer surface. The construction is shown, again diagramatically, in Figure (8). The precision bore cylinder is filamentary wood

with fiberglass in an epoxy matrix to provide the pressure containment. The pressurized gas trapped in each end

of the cylinder acts as gas springs which, in conjunction with the mass of the piston, forms a mechanically tuned

resonator. The electromagnetic and sensor design is such that the entire stator assembly is mechanically separable

from the piston-cyhnder assembly. There is no physical contact or mechanical friction in the system. Only magnetic and fluid viscous losses will occur.

LINEAR MOTOR WITH INTEGRAL MAGNETIC BEARINGS

FEATURES OF DYNAMIC BALANCER CONSTRUCTION

A

B

A. PISTON

PASSIVE FERROUS ELEMENT

CYLINDER, TRIPLE WALL, SPHERICAL ENDS

1. INNER, HERMETIC DIELECTRIC

2. ETCHED FILM DIFFERENTIAL CAPACITOR PLATES & LEADS

3 OUTER, FILAMENTARY WOUND PRESSURE CONTAINMENT

MAGNETIC SUSPENSION & LINEAR MOTOR

PERMANENT MAGNET BIASED

SERVO CONTROLLED DISPLACEMENTS

REMOVEABLE EXTERNAL ASSEMBLY

Figure 8. Features of Dynamic Balancer Construction

152
|i The entire system for the engineering test unit with associated control electronics is shown in Figure (9). No

I major changes were necessary as the result of engineering model tests. The design has been scaled up from a 4 cm

I to a 6cm diameter bore, all magnetic parts are laminated construction, and a pressure cyUnder design following the

j described approach is underway.

Figure 9. Photograph of Balancer with Breadboard Control Electronics

5. Conclusion

An active dynamic balancer is being developed which is expected to permit the use of reciprocating mechanical coolers of the Stirling cycle type on spacecraft sensors for high resolution imaging. The device has a single

moving part and non-contacting magnetic bearings and a novel Hnear motor/generator. It will provide vibration

cancellation with a reliabihty compatible with multiyear lifetime requirements of upcoming cryogenic cooHng sys- tems. In addition, an effort is being made to demonstrate advanced linear motor concepts which are applicable

to future designs. The integrated magnetic bearings are also a first and represent a potential weight savings and

complexity reduction. Non-contacting sensing techniques particularly well suited to the large area, close clearance,

requirements of the suspension and gas springs are being tested.

6. References

[1] Studer, P., "Magnetic Bearings for Spacecraft" GSFC X-721-72-56, Jan. 1972.

[2] Henrikson, et al., "Magnetically Suspended Momentum Wheels" AIAA Aerospace Sciences Conference, 1974.

[3] U.S. Patent No. 3,775,687.

[4] Gasser, M., Sherman, A., and Beale, W., "Developments Toward Achievement of a 3-5 Year Lifetime Stirling

Cycle Refrigerator for Space Applications."

153
DESIGN CONSIDERATIONS FOR MICROMINIATURE REFRIGERATORS

USING LAMINAR FLOW HEAT EXCHANGERS

W. A. Little

Physics Department, Stanford University

Stanford, CA 94305

Conventional counter flow heat exchangers are designed to operate

with the fluid streams in turbulent flow. The resultant mixing of the

fluid allows heat to be transferred from the walls of the exchanger to the body of the fluid. If, however, the dimensions of the exchanger are sufficiently small, enough heat may be transferred by conduction alone

that operation in the laminar flow regime is possible. We discuss the

design parameters of a microminiature Joule-Thomson cryogenic refrigerator operating in this regime. Such operation offers the advantages of

simplicity of design and low noise. The latter is of particular advantage for the cooling of cryogenic sensors.

Key words: Microminiature refrigerators; laminar flow heat exchanger;

Joule-Thomson refrigerators.

We have previously discussed the problem of scaling miniature cryocoolers to microminiature size [1]. As discussed by Daunt [2], the pressure gradient in a heat exchanger is related to the mass flow m of the fluid, and the diameter d of the tubing through which it

flows, by the expression

^2 „ -,5o 0.2

TT p d Re

where the Re is Reynolds number (Re = and p is the density and E, the viscosity of the

fluid. We had shown previously that for a Joule-Thomson (J-T) refrigerator with the heat

exchangers operating over the same temperature and pressure regime the diameter of the ex-

.0.5 changer tubing should be scaled so that d = ni . Where m is proportional, of course, to the

refrigeration capacity of the device. Reynolds number for the fluid flow in each part of

, 0.5 the exchanger is then proportional to m too. Thus^ as one scales the device to smaller and

smaller size^eventual ly one comes to the point where Reynolds number is too low for turbulent flow of the fluid to occur. For this reason we have considered the possiblity of designing a heat exchanger for operation under laminar flow conditions.

154
It should be noted that turbulent flow is normally considered essential for the operation of a heat exchanger for the resultant mixing of the fluid allows heat to be transferred

from the walls of the exchanger to the body of the fluid. If, however, the dimensions of the

exchanger are made sufficiently small, this mixing no longer is so important and sufficient

heat may be transferred to the fluid by conduction alone. Laminar flow operation is then

possible without loss of efficiency. In this paper we consider the conditions under which

such operation can be achieved.

In the truly laminar flow regime^one can calculate the velocity flow profile of the

fluid directly from the definition of viscosity. For a counter flow exchanger, such as that

illustrated in Figure l,the flow follows the contour illustrated in Figure 2.
As the aspect ratio o) /d for the gas channel becomes larger^ the edge effects from the

finite width of the channel becomes unimportant and the velocity contour becomes almost

independent of position. Integrating the volume flow through each channel then gives a mass flow

Again from first principles one can calculate the temperature profile knowing the

thermal conductivity of the two fluids and that of the wall separating the two streams. For

the flow profile of Figure 2 one obtains a temperature profile shown in Figure 3.

di

1 —

)

0 1—AT, ' AT,1 AT3—

1

VELOCITY AT

Figure 3. Temperature profile AT across a counter-current

heat exchanger for the flow profile on the left

Assuming that no heat enters the fluids except through this wall, and that the specific heat

per unit mass of each of the two streams is the same; as is the thennal conductivity; and,

di = da ) then one can calculate the total temperature drop AT from one side of the exchanger

to the other. One obtains

Where ki is the thermal conductivity of the fluid^Kz that of the wall, the specific heat

of the fluid, and Ti and T2 the temperatures of the hot and cold ends. This expression has

a very simple interpretation. Let us rewrite it as

156
m Cp (Ti- T2 ) = Ki 1 +

d2

d

2k:i \ 1K2/'

AT (4)

The left hand side represents the heat carried down the exchanger by the movement of the

fluid while the right hand side represents the heat conducted across the exchanger perpendicular to this flow. For exchangers of different geometry, the only difference in this ex- pression will be some numerical factors in the denominator of (4).

It is customary to define the efficiency n of a heat exchanger through the expression

Where the high and low temperatures and AT are as defined as in Figure 4 and expression (3).

Figure 4. Temperatures in the input (T) and return (T')

channels of a counter flow exchanger as a function of distance z along the exchanger

157
Using (5) in (3) one obtains the result

If we assume that the density of the gas is proportional to the pressure so p = po(p/po)

then (2) can be integrated over the length of the exchanger to give

SCALING CONSIDERATIONS

We consider now the effects of scaling such a laminar flow refrigerator. We assume the

ratio dz/di, and the thermal conductivities ki and K2 are kept constant. If we eliminate

m/o) from (6) and (7) we find that

So for refrigerators operating over the same pressure regime and having the same efficiency I should be made proportional to d^. Note that this expression is independent of

CO the width of the channels so that if we fix pi , P2 , and n then we obtain di from (8).

Using this value in (7) we can obtain w for a given mass flow. Different refrigeration ca- pacities can be obtained by varying the width of the channel with no change of the efficiency.

Consider the following example of a 25 mW J-T refrigerator operating with nitrogen gas at a boiler pressure of 3 atmospheres (87K) and an inlet pressure of 120 atmospheres at 300K.

If one consults a T-S chart for nitrogen, one finds that for a heat exchanger efficiency of

0,95 a mass flow of about 2.5 mg/sec would be required. Let 5, = 5 cm for both the high and

low pressure channels and assume '^^'^VdiKz « i (a reasonable assumption in certain configurations) then we find from (8) that di = 35 microns and from (7) the width of the channel

0) = 1.1 mm. The pressure drop in the high pressure channel will also be given by (7) and in this case would be about 0.03 atmospheres. A capillary would then be required to drop the

pressure from 120 atm to 3 atm. Again using (7) we find that this could be accomplished with

a capillary 5 cm long, 110 microns wideband 6 microns deep. For the flow through the capillary where a large pressure drop is required turbulent flow will persist for refrigerator

designs even of very small dimensions. For such cases^ the dimensions should be calculated

using the turbulent flow expression (1) rather than (2).

158
This simple analysis shows that microminiature heat exchangers operating in the laminar

flow regime can have efficiencies comparable to their larger macroscopic cousins operating

under turbulent conditions. One should also note that it is only the area which determines the efficiency of the exchanger in (6) not i alone. This is in contrast to the case for

turbulent flow where the requirement of a large Reynolds number and hence a large flow ve- locity makes it necessary for the exchanger to be long and narrow. For laminar flow^slow

flow of the fluid over a broad surface can be used without loss of efficiency. This simplifies the design of miniature laminar flow exchangers and reduces the dimensional constraints

in their fabrication. In addition, the absence of turbulence in the exchanger should reduce

mechanical noise in the device. The mi crofabri cation techniques described in the accompanying paper by Hollman and Little [3] together with a laminar flow design should make it

possible to build efficient J-T refrigerators an order of magnitude smaller than those described hitherto.

Support for this work was provided by the Office of Naval Research Contract N00014-78-

C-0514.

REFERENCES

[1] Little, W. A., "Scaling of Miniature Cryocoolers to Microminiature Size"

Proc. NBS Cryocooler Conf.

Ed. J. E. Zimmerman and T. M. Flynn, Special Publication 508 (April 1978)

[2] Daunt, J. G. , Encyclopedia of Physics, Ed.S. FlUgge, Vol XIV, Low Temperature

Physics I, Springer Verlag, Berlin (1956).

[3] Hollman, R. , and Little W. A., "Progress in the Development of Microminiature

Refrigerators using Photolithographic Fabrication Techniques", this issue.

159
PROGRESS IN THE DEVELOPMENT OF MICROMINIATURE

REFRIGERATORS USING PHOTOLITHOGRAPHIC

FABRICATION TECHNIQUES

R. Hollman and W. A. Little

Physics Department, Stanford University

Stanford, CA 94305

We report on progress in the development of fabrication techniques

for the construction of microminiature cryogenic Joule-Thomson refrigerators using photolithographic means. A new technique for the fabrication

of the devices from glass will be described. Consideration has been

given to the design of refrigerators operating in this 50 - 200 mWatt

region at 77K and 20 mWatt below 20K. Recent results on their operation

will be reported.

Key words: Cryogenic; Joule-Thomson; photolithographic fabrication;

microminiature devices; refrigerators.

In 1977 Little [1] discussed the scaling laws involved in scaling miniature cryocoolers

to microminiature size. It was suggested further "that the fabrication of microminiature devices might be accomplished using photolithography. Preliminary work on the development of

such a technique for the fabrication of a Joule-Thomson (J-T) refrigerator on a silicon

wafer was reported at a conference in Charlottesville in 1978 [2]. We wish to report here

the development of a class of J-T liquefiers designed to operate at liquid nitrogen or hydrc

gen temperatures with a capacity of the order of 100 mW. These have been fabricated using

photolithography on glass. Single stage devices have been operated successfully below the

critical point of nitrogen with a refrigeration capacity of 100 mW at 122K. Typical channel

dimensions are of the order of 100 microns, and the flow rate about 30 ml/sec (STP) of nitre

gen at an inlet pressure of 1600 lbs/in^. The theory of turbulent flow heat exchangers as treated by Daunt [3] was found to be applicable on this scale and our results at these small

flow rates confirm the belief that efficient heat exchangers can be built in a very small

space. The principal problems in the construction of these devices are now understood. We expect these one stage and two or three stage refrigerators of this type will be useful in

cooling very small material samples, or instruments in which heat production is minimal.

A J-T refrigerator consists of three sections: a high pressure inlet channel, a low

pressure outlet channel which together form the heat exchanger; and, a narrow capillary con- necting the two to confine the pressure drop to the cold end (Fig. 1).

(a) (b) (c)

Figure 1. Detail of microminiature J-T refrigerator

showing

(a) inlet and outlet ports

(b) heat exchanger

(c) capillary expansion channel

160
I The present devices are constructed by etching grooves in a glass surface, then sealing

them with a glass cover plate. + Earlier work [2] used methods described by Terry [4] to etch

channels in silicon sealed with pyrex. These silicon devices operated successfully with

I carbon dioxide or ethylene as the refrigerant. However, attempts to operate them to LN2

temperatures failed because the high thermal conductivity of Si necessitated the design of

excessively fragile heat exchangers which could not contain the high pressures required for

efficient operation with N2. For these larger temperature differentials, the low thermal

conductivity of an amorphous material such as glass is necessary. The geometry of the devices

together with the exacting sealing criteria required the development of an etchning technique

[which would allow deep etching if the grooves with minimal lateral etching or undercutting.

Chemical etching with HF proved unsatisfactory. A new technique for the precise, directive etching of any amorphous or crystalline material was developed.''' The surface to be

etched is prepared by coating witji a thick water solution of gelatin activated with potassium

dichromate. It is then dried, exposed through a mask to ultra violet light and developed in

hot water. The unexposed gelatin washes out leaving a pattern of unprotected glass. This

resist forms an exceedingly tough thin protective coating. The coated glass is then mounted

i on a rotating holder in a lathe-like device and abraded with 27 micron AI2O3 powder entrained

j in a high velocity gas stream using an S.S. White Airbasive unit with the nozzle scanned

I across the surface of the glass. Channels of precisely controlled depth in the range of

I 2- 100 microns can be etched in this manner.

To complete the fabrication of the refrigerators a cover slide is sealed to the etched

plate using a thin film of a UV-curing adhesive. Stainless steel hypodermic tubing epoxied

to the glass" plate provide inlet and outlet connections (Fig. 2). Details of the etched

slide are shown in Figure 3 and Figure 4.

(a) (b) (c) (d)

Figure 2. Steps in fabrication of microminiature

refrigerator

(a) coated, exposed and developed glass slide

(b) slide after abrasive etching

(c) slide with cover plate bonded to it and

inlet and outlet ports drilled

(d) completed refrigerator with stainless

steel gas lines epoxied to inlet and

outlet ports

161
The planar construction together with the need for containing the high pressure requires

a cross section of material much greater than that of the channels themselves. For this reason, solid conduction is a significant heat load. This dictates the use of glass for

construction, and makes a thin, narrow device preferable.

Bonding the glass proved to be difficult. An adhesive is needed which will hold the

pressure used with a layer less than 10 microns thick. Complete wetting of the surfaces

must be obtained without filling the narrowest channels with adhesive.

The purity of the refrigerant gas is critical for these open-cycle devices. The capillary can be blocked by about 1 microgram of condensed impurities. We have used nitrogen of

99.996% purity, and passed this through an activated charcoal trap at ambient temperature.

Operation for 1 to 2 hours has been achieved to date. Longer operation will require higher

purity gas or a trap built in to the heat exchanger. The limitation on the lowest temperature reached is the back pressure of the gas in the return channels-. The present devices were designed to have a rather deep capillary to minimize difficulties with the adhesive

during the bonding step. This results in a larger flow and a back pressure at the boiler of

the order of 200 lbs/in^, limiting the low temperature end to llOK or above. A minor change

in the design should correct this.

Figure 3. Detail of etched channels of heat exchangers

showing high pressure channel between the two low pressure return channels. The high pressure

channel is 300 microns wide and the separating

walls 150 microns wide.

162
Figure 4, Detail of etched slide showing capillary

expansion channel and part of boiler.

Capillary is approximately 100 microns

wide and 50 microns deep.

The design currently in use measures 75x12x2 mm and has a 30 cm long heat exchanger.

The cooldown time is 8 minutes. Smaller, less massive devices are now being developed using

thinner glass substrates which should result in substantially faster cooldown times. A hydrogen liquefier with integrated nitrogen precooler is also being developed for operation

to 23K.

The present design is being considered for use in microcalorimetry , optical reflectance

spectroscopy, and IR transmission spectroscopy experiments. A sample mount can easily be

integrated into the design. For reflectance spectroscopy, the refrigerator and its vacuum

shroud are small enough for mounting on a goniometer at room temperature.

This work was supported in part by the Office of Naval Research, Contract N00014-78-

C-0514.

REFERENCES

[1] Little, W. A., "Scaling of Miniature Cryocoolers to Microminiature Size"

Proc. NBS Cryocooler Conf.

Ed J. E. Zimmerman and T. M. Flynn, Special Publication 508 (April, 1978).

[2] Little, W. A,, "Design and Construction of Microminiature Cryogenic Refrigerators"

in Future Trends in Superconductive Electronics,

Ed B. S. Deaver et al , APS Conf. Proc. 44, 421 (1978).

[3] Daunt, J. G. , Encyclopedia of Physics, Ed. S. FlUgge, Vol XIV, Low Temperature

Physics I, Springer Verlag Berlin (1956).

[4] Terry, S. C, Ph.D. thesis Stanford University (1975) (NASA Technical Report

4603-1, 1975).

t Patent Pending.

163
LOW TEMPERATURE REGENERATORS

Lawrence A. Wade

Lockheed Palo Alto Research Laboratory*

3251 Hanover Street, Palo Alto, California 94304

A study was conducted to examine ways of enhancing the heat capacity

of regenerators used in refrigerators for achieving refrigeration below

30°K. It was found that, due to an increase in the energy of their surface

vibrational modes, thin films, thin wires,and ultrafine particles will ex- hibit a higher specific heat than that which would be normally expected for

a bulk sample of the same material.

Based on an idealized refrigerator performance model, it is predicted

that a Stirling cycle refrigerator having the third stage of its regenerat- or replaced with an ultrafine particle (22 8 dia. lead particles) matrix

whose void volume is partially filled with mercury, will deliver at a cooling temperature of 10°K, 2.75 times as much refrigeration as one with

a typical third stage regenerator made up of 0.005" dia. lead spheres.

A. table showing the expected improvements in refrigerator performance

from the use of several ultrafine particle configurations is presented.

Key words: Heat capacity; regenerators; specific heat; refrigeration;

cryogenics

.

1 . Specific heat enhancement

The efficiency of refrigerators for cooling below 20°K has been, to date, severely

limited by poor regenerator performance. The primary cause of poor regenerator performance

at these temperatures is the fall -off of the regenerator matrix's heat capacity.

This paper deals with how the heat capacity of the regenerator might be improved.

The model typically used to describe a material's specific heat at constant volume is:

if Author's present affiliation. This paper presents work partially done while with Scitech.

Corp. It was supported by AFWAL Flight Dynamics Laboratory, Wright-Patterson AF Base, Ohio,

under Contract F3361 5-78-C-3425. Also presented is work done at Lockheed Palo Alto Research

Laboratory under IR & D funding.

164
0)

This model is based on the assumption that the material is an elastic continuum. In the case of extremely large surface to volume ratios this assumption is not accurate. The

number of lattice sites and the number of possible vibration modes were significantly undercounted. E. W. Montroll discerned this and derived a set of frequency distribution functions taking this into account. Using the same approach and assumptions as Debye, with the

exception already noted, he derived a new specific heat model.

= 3T^ + aT^ + Y(l+e)T , (2)

where

S'p^/^ M^^^ a «

and

e

\^ } 24 ^n*^

where S' is the specific surface area, p is density, M is atomic weight, 9^ is the Debye

temperature, is Avogradros number, and n* is the average number of conduction electrons per

lattice site. The surface contribution would be expected to become significant only at very

low temperature and for large surface to volume ratios. The details of this derivation are beyond the scope of this paper, but they may be found in Reference II J.

A series of measurements of the specific heat of ultrafine particles have been made

over the years. They have been performed on metals such as palladium and vanadium, as well

as dielectrics, such as magnesium oxide. They indicate that Montroll underestimated the

surface contribution by a factor of between 4 and 6 (Ref. [Ij).

Modifying the corrections made by Montroll to take the experimental results into

account, we find

Cy = BT^ + 4aT^ + Y(l+4e)T . (3)

Applying equation (3) to lead and mercury, it is evident that the surface effect is

indeed quite significant (see figures 1 and 2).

It was assumed in this study that the assumption from which equations (1) - (3) are

derived (that T « 6^) was accurate up until T = Qq/S. Thereafter it was assumed that the

surface enhancement effect would decrease to 0 at 9^/3. No measurement of ultrafine

particle specific heat has been attempted yet in the range where the assumption T « 9^ is not reasonably valid.

165
0 50 150 0 50 100 ISO

Figure 1. Lead volumetric specific heat. Figure 2. Mercury volumetric specific

heat

There are a number of different lattice dynamics models which have been developed since

the work by Montroll. However, to date, none of them appear to be superior to the simple

model of eq (2) in predictive value. For this reason, the Montroll model was modified to fit the existing experimental data (eq (3)).

An experimental program examining the specific heat of ultrafine particles up to 100°K

could give some insight into the physics of specific heat which would be valuable to future

workers in this field, A program of this type would allow an empirical modification to be

made on eq (2) which would improve on the accuracy of the predictions made by eq (3). In addition, the assumption made as to the degradation of the surface contribution, as a function of temperature, could be eliminated and instead, an empirically based model used to predict refrigerator performance.

166
2. Utilization of ultrafine particles

There are two candidate methods for the utilization of ultrafine particles, both of

which are derived from the powdered metallurgy industry. The first method is derived from

techniques used during the formation of sintered parts.

2.1 Green compacts

The first stage of preparation of sintered parts involves pressing metal powder into

"green compact". At this stage the metal will hold together but the surface welding between

particles is minimal. The extent to which the surface welding occurs is related to the pressure exerted during compaction. The second stage involves heating the green compact to just

under the materials melting point. At this temperature heavy surface welding takes place,

"sintering" the powder together, and thereby increasing its mechanical strength. For this

application, heavy surface welding would be undesirable as it would destroy the discreteness

of the particles. An optimum pressure must be determined which will provide maximum

strength and density of the element without causing an excessive amount of surface welding.

Once formed into a green compact, gas flow passages will need to be fabricated in order

to minimize the pressure drop through the regenerator. The end result would approximate a conventional screen. If the gas flow passages are aligned, pressure drop losses could be

kept to a minimum.

It is also expected that the helium gas used as the working fluid will contribute to the regenerators heat capacity. In the model used when predicting the performance of a refrigerator with an ultrafine particle regenerator, the helium gas was at a base cycle

pressure of 40 atm. It would seem reasonable that some of the helium gas would be adsorbed and absorbed, and would thereby contribute to the regenerator's heat capacity. In the case of 22^ dia. particles this effect is expected to be maximized because the particles

are only about 5 lattice sites across and have a large specific surface area C200 ra /g).

Absorption refers to the diffusion of the gas through the host metal 's lattice structure, thereby forming, a solution. Gas which has been absorbed into the lattice can be

considered to represent, from the standpoint of the gas specific heat, a solid. Unfortunately,

it is found that helium will not diffuse through a metal lattice structure (Ref. [2]). Therefore, there will be no contribution to the regenerator matrix's heat capacity from the

absorption of helium gas.

Adsorption is caused by an attractive force which is exerted in a direction normal to the surface by the atoms at the surface of a solid. The attractive force is a result of

Van der Waals forces which arise from dipole interactions between the solid and the gas.

167
This is referred to as physisorption, and is the only kind of adsorption which can occur

between an inert gas and a metal. It is therefore found that at a gas/solid interface

the concentration of gas exceeds that of the gas not in the solid's immediate vicinity. The

excess gas at the surface is referred to as having been adsorbed. At the warm end of the

regenerator (35°K), it is reasonable to expect that only a few percent of a monolayer of gas

will be adsorbed. At the cold end (10°K), about one monolayer will be adsorbed (Ref. [3J,

[4]). In terms of the specific heat contribution from the adsorbed gas, it is best to consider it to be a liquid.

A further contribution to the regenerators heat capacity comes from the helium trapped

in the pore structure of the green compact which is not in the gas flow stream.

Green compacts do have the disadvantage of low mechanical strength. It is uncertain

what the long term effect of gas cycling would be on a green compact. As it is not presently possible to contain particles smaller than 200 %, if any particle should break off the

compact some migration would result. This is undesirable, as it is conceivable that seal

and bearing deterioration could result.

2.2 Mercury impregnated green compacts

The second method of utilization also involves the fabrication of a green compact;

after which mercury is pressed into the pore structure. A significant portion of the mercury (greater than 50% if the process is repeated) will be retained in the green compact.

This results in the matrix having a significantly higher heat capacity. Mercury has a higher low temperature specific heat than lead, a higher density, and a lower Debye tem- perature. This results in a large surface contribution to the specific heat. This process

is referred to as mercury porosimetry.

This approach has several advantages. Pressed into the pore structure, the mercury

will have a high surface to volume ratio, and it is hoped will thereby evidence a surface

enhancement of its specific heat. The thermal conduction of the compact is also expected

to be enhanced. This will result in a more efficient heat transfer between the regenerator

and the refrigerant gas. The contribution to the regenerator's specific heat by the mercury

is the most important advantage. Because of the properties already stated, mercury would be

the preferred regenerator material if it weren't a liquid at room temperature; but, by

containing the mercury within the ultrafine particle pore structure, this difficulty is overcome.

It is expected that some amalgamation will occur between the mercury and the host

metal, which will result in a much stronger structure. This results in the individual particles being more tightly bound to the compact structure and will probably solve the

potential problem of particle migration. It is unknown whether a sufficient boundary will

168
exist after amalgamation to form a wave reflection, or scrambling boundary, for the low

frequency acoustic vibrations, and therefore, whether the full Montroll effect will be

real ized.

3. Anomaly materials

An alternative to using ultrafine particles would be to find a material with a very

high bulk specific heat. The best materials which has so far been found is GdRh. It evidences -a very high heat capacity compared with lead below 20K. There would appear to be at this time no reason why GdRh could not be formed into .005" dia. spheres or into 400 mesh

screen.

Because GdRh has a relatively high Debye temperature; in calculating the specific heat

of GdRh from equation (3) it is found that the surface contribution is only about 3% of

its total specific heat. Therefore, in this report, it is examined as having been fabricated

into .005" dia. spheres.

GdRh is exceptional in that it evidences a broad band specific heat anomaly extending

from 20°K to below 2°K (Ref. [5J). Most materials which evidence an enhanced specific

heat do so over a narrow range of temperatures; or typically only a few degrees. Once the

temperature profiles of a regenerator are better understood, another approach might be to utilize a number of materials which evidence this sort of enhancement (called a lamda

transition-normally caused by a phase transition) in a segmented approach to regenerator

design. Examples of, this type of material are DyRh^B^ and (ErHo)Rh^By.

4. Regenerator analysis

The refrigerator performance model is based on an idealized 3-stage Stirling cycle

design. The refrigerator is assumed to have a base cycle pressure of 40 atm and to operate

at 600 rpm. The third stage is designed to provide .3 watts of net refrigeration at 10°K

with a gross output of 2.785 watts utilizing .005" dia. lead spheres as the regenerator

matrix. This will require a 167 design NTU value (A) for the regenerator, and a refrigerant gas mass flow rate of .8077 g/sec. The NTU value is defined as

where Aj is the heat transfer area, U^^ is the overall coefficient of heat transfer and

C = ific is the gas flow stream capacity rate. The refrigerator performance model is desy r cribed in detail in Ref. [1]. The regenerator efficiency was determined using a computer

program based on work by Lambertson (Ref. [6]). The program uses the finite difference

method to calculate the regenerator's temperature distribution and efficiency. It is assumed that there is no outgassing of the helium contained in the regenerator's pores

due to pressure fluctuations.

169
The results described are functions of regenerator heat capacity only. All else,

including pressure drop (7 psi assumed), is held constant for all the cases.

The cases described utilize the .005" dia. lead spheres regenerator matrix as the

baseline. Case 2 utilizes .005" dia. GdRh spheres in the last third of the regenerator,

the first two-thirds (warm end) being made up of lead spheres. Case 3 deals with 22 %

lead particles in a green compact without gas contributions taken into account. Case 4

is the same as Case 3 except that gas contributions to the regenerator's heat capacity are

taken into account as described in section 2.1. Case 5 is a mercury impregnated 22 8 lead

particle regenerator without amalgamation. Case 6 is the same as Case 5 except that it is

assumed that 75% of the surface specific heat contribution is lost to amalgamation.

The data shown is the percent of required refrigeration delivered where the required

refrigeration is 0.3 watts at 10°K, the design NTU value (X), the effective NTU value (X^),

where

0 1-e

and the regenerator efficiency (e). Note that X^ and e are the values for the third stage

of the regenerator, not for the entire regenerator.

From the results presented in Table 1 it can be concluded that:

(1) GdRh substituted for lead in the cold end of a regenerator

gives a significant improvement in refrigeration over a lead sphere

regenerator.

(2) Although a significant portion of the 22 % green compact regenerators heat

capacity is contributed by gas adsorption and gas in the interstitial pores;

lead ultrafines, even without the gas contributions, have a high enough

specific heat to give significant improvements in regenerator efficiency over conventional regenerators. It can also be seen that if lead proved

to be unsuitable for green compact fabrication, a material with a lower

specific heat might be utilized in the ultrafine particle configurations,

and by taking advantage of the gas contributions, still significantly improve

regenerator efficiency.

(3) That lead ultrafine particle green compacts which are impregnated with

mercury offer the greatest potential gains in refrigeration.

(4) Although amalgamation might reduce the surface contribution to the regenerator's heat capacity, the mercury bulk specific heat is large enough to provide significant improvements in refrigerator performance. It is apparent

that maximizing the quantity of mercury contained within the small particle

pore structure will result in optimum regenerator efficiency.

170
TABLE 1. A comparison of different regenerator configurations

tase iNO. Description

% of Required

Ketri geration

Del ivered A ^0 e

1 .005" dia. lead spheres 100% 167 54.6 .9820

2 GdRh spheres + lead spheres 193% 167 74.2 .9867

•3 C,Cr\ 1 cuU I cell UUMipaUL 1 fJ

4 Case 3 + gas effects 253% 167 96.1 .9897

5 Mercury impregnated 22^ lead

green compact

275% 167 107.7 .9908

6 Case 5 + amalgamation base 271% 167 105.4 .9906

5. Conclusions

An enhancement of materials specific heat can be accomplished by the use of ultrafine

particles of the same material. The utilization of ultrafine particles presents a number of

difficult problems; however, the improvement in refrigerator performance appears to be significant enough to merit experimental research to determine the mechanical properties of

green compacts in different configurations and to make an accurate determination of their

heat capacity over a wide range of temperatures. This will allow more accurate regenerator

efficiency predictions to be made and the feasibility of utilization analyzed. Fabrication

of GdRh into .005" dia. spheres would also seem merited as it presents a relatively simple

solution to immediate refrigerator needs.

Part of this work was sponsored by Flight Dynamics Laboratory,

A.F. Wright Aeronautical Laboratories, Wright-Patterson Air Force Base,

Ohio, under contract F33615-78-C-3425, Project 2402, Task 240204,

"Aerospace Vehicle Environmental Control," Mr. Ronald White was the

Project Engineer for AFWAL/FIEE. This work was conducted at Scitech

Corporation, Santa Ana, Ca. The project manager was Mr. Henry Yoshiraoto and the principal investigator was Mr. P. J. Walsh.

The extensions on this work were carried out at Lockheed Palo

Alto Research Laboratory, Palo Alto, Ca. and sponsored through Internal

Research funds.

I would especially like to thank Mr. Jerry Walsh for the help and

guidance he gave me throughout this entire project.

171
6. References

j

[1] P. J. Walsh, Low Temperature Regenerator Study, Interim Report , Air Force Flight Dynamics'

Laboratory, Wright-Patterson Air Force Base, Ohio, 45433, AFFDL-TR-79-3099, Aug. 1979,

| AO 80892.

i

[2] R. M. Barrer, Diffusion in and Through Solids , Cambridge University Press, 1973, Chapter:

3-4.

I

[3] R. C. Tompkins, Chemisorption of Gases on Metals , Academic Press, 1978, Chapters 1 and 3.'

[4] J. H. de Boer, The Dynamical Character of Adsorption , Oxford, 1953, Chapters 1-4. i

I

[5] K. H. J. Buschow, J. F. Olijoek, and A. R. Miedema, "Extremely Large Heat Capacities 1 between 4 and lOK," Cryogenics 15, 261 (1975). j

[6] T. J. Lambertson, "Performance Factors of a Periodic-Flow Heat Exchanger," Trans. ASME

80, 586 (1958).

I

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1

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172 '

I'

i
MEASUREMENT OF THERMAL PROPERTIES OF CRYOCOOLER 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 'V' 0.5 mW/cm K and a specific heat c 'v- 25 mJ/cm K at 10 K. Thermal equilibrium times will of course be proportionately greaPer

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, th^^ 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.

173
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 is a uniform semi -infinite bar with a heater on the free end, x = 0,

and thermometers to measure temperatures T, and at two points x, and Xp (Fig. 1). The

simplest initial condition is to let the b^r 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-dimensional diffusion

equation for this situation is

3x2

_3T

3t

(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/ukC

/t e CxV(4l<t)

kA ^

^'^'^ Mkt^ .

(2)

where erfc means the complementary error function, and A is the cross-sectional area. This

solution is plotted as a function of x and of t in Fig. 2,

HEATER

THERMOMETERS^

T2 ^

3

Fig. 1. Experimental arrangement for thermal properties measurement. The heater is turned on abruptly at t = 0 (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 is suspended in a vacuum and

surrounded by a liquid he! ium°cooled shroud to minimize heat leaks to it. The

leads from the carbon resistance thermometers are heat sunk to this shroud for

simi lar reasons.

174

I'
T(tp() - T„

T(t,o)-To

(a)

0 0.5 1.0 1.5

DISTANCE, x/{2Vkt/C)

-^I^ (T(t^)-T„)

xP
fim. ^"^.I'f^?^ °^ measurement is to record the temperatures, T, and as a function of

. T h ' P"'" of curves of the shape of Fig. 2b, but^with different time scales

s nee T] changes more rapidly after the power is turned on than T,, and of course To is always less than T.,

. From these curves we can read off a pair of^numbers T, and To^for a particular time t . For that particular time we then have both coordinates ' (T^ , x?) and (Tg^.x^) for two points on the curve of Fig. 2a, which establishes the horizontal aM

vertical scales of that curve. This process involves use of the ratios x,/Xp and T,/Tp 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 = 0 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 Tp at some convenient time t. How- ever, measuring T, and 1^ 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 To do not rise very much above the

initial value Tq. 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 (the position at which To is measured) at about the middle of the real bar,

and then choose t' such that the rise in To 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 iurve of Fig. 2b might indicate that one or more of these

conditions has been violated.

6.0 I4.5 I I \ I I 1 1 0 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 crosssectional area of 0.895 cm . 176
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 regions.

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

[1]. 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).

177
GAS HEAT SWITCHES

Emanuel Tward

Jet Propulsion Laboratory

4800 Oak Grove Drive

Pasadena, California 91103

Thermally operated gas heat switches for use in refrigeration

systems operating in the temperature range 4K to 300K have been designed and constructed. The switches are operated by a miniature adsorption pump which when heated supplies gas to the switch, turning

it on, and when cooled removes gas from the switch, turning it off.

The switches have been operated at 4K, 77K, and 300K with He or H^^ gas adsorbed on either charcoal or zeolite 13X. Switching times as low as a second and switching ratios of 80 have been achieved in

miniature devices whose total mass is 8 grams.

Key words: Cryogenics; gas adsorption; heat switch; refrigerators.

1 . Introduction

In order to cool sensors in spacecraft, refrigeration systems

covering a broad range of temperatures from less than IK to ambient

and having service lifetimes of up to 10 years are being designed and

built.

Some of these systems require thermal switches both for switching

the load to the refrigerator and as an essential component of the re- frigerators themselves. Since the primary goal being addressed is one of lifetime and reliability with the potential for miniaturization,

it was decided to develop a gas heat switch. Since the switching

function is accomplished by changing the gas pressure between two surfaces, a gas supply and a pump are required. Both of these functions

are provided by an adsorption pump connected to the switch. This

switch, together with its adsorption pump operator, form a permanently

sealed system with no moving parts. This permanent sealing, to- gether with the absence of mechanical motion, should provide the re- liability and long lifetime which are required.

178
2. Principl e of Operation

The thermal conductance between two parallel surfaces of area A at temperatures and separated by a distance d depends on the gas

pressure between the surfaces (1,2). At "high" pressures where the mean free path L <<d (classical conduction region) the thermal con- ductivity of the gas is essentially independent of pressure and varies

almost linearly with temperature (3) as shown in Figure 1. At

"low" pressures where Lp»d the thermal conductivity of the gas is

given by (1

)

It + il

I RI ^ I p \^

where e is an accommodation coefficient for the surfaces, 7= Cp/Cv

is the ratio of specific heats for the gas, R is the gas constant, P

is the pressure, M is the molecular weight of the gas, and T is the

temperature. It is therefore apparent that at some fixed operating

temperature T the thermal conductance between two surfaces may

be varied from some high fixed value in the classical regime to some

lower value which is pressure dependant. In a real switch, however,

the minimum conductance is determined by other thermal conductances

which are operating in parallel to the conductance through the gas.

These conductances arise due to the heat transfer by radiation between the two surfaces (usually negligible in our applications) and

due to the heat transfer by conductance through the supporting

structure holding the two surfaces apart. A schematic of a design

for a real heat switch is shown in Figure 2. Two high conductivity

plates (e.g. copper) of area A are separated by a small gap d and

held apart by a low conductivity support (e.g. thin wall stainless

steel tube). The gas gap is connected via a stainless steel capillary to a small adsorption pump. The thermal conductance between

the two plates at the different temperatures T-j and T^ can be modeled

as in Figure 3 (a), where the plates 1 and 2 have thermal resistances R-j and Rp, respectively, the isolation support has a thermal resistance R

^ s

,

the radiation resistance is R^,and the thermal resistance due to changing the gas pressure p is R(p). In practice for the devices

which we have made R^ , R2«R5 and R^»R(p), R^ and therefore the equivalent

circuit is adequately described in Figure 3 (b). Therefore, if the

gas pressure p is changed from some value P^^-g^ to some value p-j^^

where R (piqw) ^^'^s' switch ratio is given as

179
LOW CONDUCTIVITY

SUPPORT

HIGH CONDUCTIVITY PLATE

GAP

AREA A

CAPILLARY

ADSORBER

THERMAL RESISTANCE

R^ - PLATE 1 RESISTANCE

R - ISOLATION SUPPORT

^ RESISTANCE

R - RADIATION

^ RESISTANCE

R- ^ - VARIABLE PRESSURE

RESISTANCE

Rg - PLATE 2 RESISTANCE

Figure 2. Heat Switch.

180
R VVAr

AAAr -VVAr -OT,

(a)

(b)

Figure 3. (a) Model of Heat Switch resistances,

(b) Simplified Heat Switch model.

181
R'off R

s

Ron Phigh

3. Adsorption Pump

The adsorption pump acts as a source and sink for the gas. It

is operated by heating and cooling the adsorber which causes gas to be either released or readsorbed respectively (4,5), A number of

materials can be used as adsorbers. In a heat switch where the thermal

conductivity of the gas is important, the most appropriate gases for use are He and (Figure 1). These gases may be adsorbed onto charcoal or zeolite for example, or in the case of H2 on to metal hydrides. An important feature of these adsorbers is that the adsorbers have ex- tremely large effective surface areas per unit volume ( ~500 m /gm)

and hence can store liquid densities of the gas at .low pressures.

3 Storage capacities of 200 cm STP/gm adsorber are possible. If the

temperature is raised, gas is released and the pressure builds up. Since a large amount of gas is stored on the adsorber and the pres- sure range of operation is usually in the millitorr to few torr range,

therefore orders of magnitude pressure changes are possible with small

temperature changes and with small amounts of adsorber. The principle

of operation of the switch operator (adsorption chamber) is shown in

Figure 4.

TO SWITCH

HEAT IN

Figure 4. Heat Switch Operator. The adsorption pump is heated to turn

the switch on and cooled to turn the switch off.

182
When the heater is off, the adsorber is held at the temperature T^.

At that temperature, the bulk of the gas is adsorbed and the pressure

has been adjusted to be P-jQ^j sufficiently small that the conductance

between the plates is determined by the isolation support in the switch

When the adsorber is heated by the heater the pressure rises to a value Pf^^-gp, where the mean free path Lp«d (the gap). For example,

using He in a switch operating at 4K, Lp [a|^] is 10" "^mm at 1 torr. _2

Therefore, for a gap larger than, say 10 mm, a pressure of 1 torr is

large enough for P^^^-g^-

The quantity of adsorber and the power required to operate the

switch are a function of the temperature T^ at which the switch is

operated. This results from the strong dependance of the adsorption

capacity of the materials on temperature. If T^ is near the boiling

point of the operating gas, then very small quantities of the adsorbers

such as charcoal and zeolite are required because of the large sto- rage capacities of the adsorbers. For example, one of the switches

-1 3 which we have made has a gas volume of approximately 10 cm , At

4.2K zeolite 13X adsorbs ~1 50 cm"^STP/gm at 10 torr (6). To pro-

-4 3 -4

duce 1 torr in this volume requires only 10 cm STP of gas or 10

-2

of the gas stored on 10 grams of zeolite.

The power required to operate the switch is also a function of

the switching time required. For rapid turn off time the thermal

resistance of the thermal link must be reduced, thus causing an increased heat load at the temperature T^ and increased power re- quirement to turn the switch on.

4. Measurements

Measurements were made in a miniature heat switch similar to that shown in Figure 2, The switch is i^" diameter x .200" long and con

tains a gap of .001". The isolation support is a stainless steel

tube^i" diameter x .001" wall. The plates are made of copper. Total mass of the switch is 8 grams.

Figure 5 shows the response of the switch to the alternate

heating and cooling of the charcoal pump containing a few mgms of

charcoal. The data was taken with one side of the switch (T-j ) fixed

at 4.2K. 6.5 mW is applied continuously to the lower plate of the

switch carrying the increased temperature 1^ recorded in Figure 5. In this example, is relatively small so that the switch off time

can be made fast. The He working gas is adsorbed on the activated

183
Figure 5. Response of Heat Switch to operation of adsorption pump.

charcoal at T^=4.2K. At t=0 seconds the 6.5 mW is being applied to the bottom of the heat switch while the switch is non-conducting.

At t=110 seconds the charcoal pump is heated. Within 1 second the

switch is conducting. At t=l 50 seconds the charcoal pump heater is turned off. Within 1 second the bottom of the heat switch begins

warming indicating that the switch has turned off. The switch ratios

for this switch were also measured at 77K. With He as the working gas, this switch has a resistance while on R„„=16 K/W. R ^^=1160K/W

on off

giving a switch ratio of 72.5. With H2 as the working gas, R^^=15K/W

giving a switch ratio of 77. The research described in this paper was carried out at the

Jet Propulsion Laboratory, California Institute of Technology, under

NASA Contract NAS7-100.

184
5. References

R.R. Conte, Elements de Cryogenie, pp. 17-24, Masson et Cie,

Paris (1 970).

R.J. Corrucinni, Vacuum , Vol. VII & VIII, 19 (1959).

G.K. White, Experimental Techniques in Low Temperature

Physics, p. 82, Oxford U. Press, Oxford (1968).

R.R. Conte, Elements de Cryogenie, p. 68, Masson et Cie,

Paris (1970).

A.C. . Rose-Innes and J.J. Rowland, Journal of Physics E: Scientific Instruments, 5, 939 (1972).

J.G. Daunt and C.F. Rosen, Journal Low Temp. Rhys., 3,

89 (1970).

185
OPERATION OF A PRACTICAL SQUID GRADIOMETER

IN A LOW-POWER STIRLING CRYOCOOLER+

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

I system supported by a split Stirling refrigerator. The I magnetic interference generated by the cryocooler was I sufficiently small to permit the acquisition (for demon-

| stration purposes) of a reasonable magnetocardiogram. I Preliminary investigation of the source of the residual ; refrigerator interference indicates that cyclic temperature variation is the largest single factor influencing I the gradiometer output. A simple electronic temperature ! controller has been added to a different cryocooler to I study temperature stabilization. Better than 1 mK

j stability has been achieved at 8.5 K. I

I Key words: Cryocoolers; low temperature; magnetic j gradiometer, refrigeration; SQUID; Stirling cycle; I superconductivity. i

j

I 1. Introduction I

I The economic viability of superconducting instrumentation would be greatly enhanced if | properly matched cryocoolers were readily available to support their operation. By properly i 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), j 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 magnetoj meters and magnetic gradiometers which are designed to sense environmental magnetic fields ! or gradients (e.g., geophysical -.fields or biological fields). The SQUID magnetometer has a sensitivity of the order of 10" times the earth's magnetic field and thus extremely small

j angular vibrations translate directly into interference if the device is operated in such a

j field. In this paper we discuss this and other potential interference sources and describe i 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 i,

r The cryocooler used in this work has been recently described by Zimmerman [2]. Figure f 1, taken from this reference, shows the details of construction of the cold end of this I 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

j just over 2 atmospheres (gauge pressure). The displacer stroke is 0.7 cm.

!

t Contribution of the U. S. Government, not subject to copyright. Work supported by the

| Navy Coastal Systems Laboratory. i tt Present address: Engineering Department, University of Utah, Salt Lake City, Utah. '

186
Displacer

Push Rod

Pinned

Displacer

Joint

Aluminum

Flange

9.5—

Vaccum

Wail

4.8

Annular

Expansion

Space

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

Vacuum Wall

Fig. 1. Details of 5-stage, split Stirling cryocooler. Dimensions are in millimeters.

Figure taken from reference [2].

187
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 be- tween 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.

Fig. 2. SQUID gradiometer. The SQUID is inside a niobium cylinder.

188
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

jflux quantum with a noise level of 10" T/m in a 40 Hz bandwidth. For thermal anchoring,

I 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 fsee 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.

DISPLACER AND CYLINDER

IN VACUUM JACKET

ALUMINUM ROOM

Fig. 3. Mounting of the cryocooler in an aluminum shielded room.

189
4. Potential Sources of Interference

In previous papers we emphasized the refrigeration aspects of the cryocooler while alluding to certain charactertstics 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 Itttle 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 fielium 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 l^Hz 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 wel 1

.

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 direct 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.

190
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

IT X (2.4)^ cm-^. For nylon with a susceptibility of 10"^ in a field of 5 x 10"^ 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 likely 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 cur- rent 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 IQ-I ' 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

191
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 is somewhat dependent on temperature since the slope of the I vs. T curve is a function of T and some dependence is thus expected. ^

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 lO-H 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 used 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 canceledthe 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 is probably a result of the rigid mounting to the shielded room (see Section 3).
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.

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.

BRASS BLOCK

193
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).

194
CURRENT STATUS OF HIGH TEMPERATURE JOSEPHSON DEVICE TECHNOLOGY

M. Nisenoff

Code 6854, Naval Research Laboratory, Washington, D.C. 20375

The fabrication of Josephson devices from refractory

thin films with high superconducting transition temperatures (T ) should yield devices which are rugged and

reliable, and capable of withstanding repeated thermal

cyclings between room and cryogenic operating temperatures. If the onset of Josephson response could be

raised to above 15K, the devices could be operated near

lOK at a reduced temperature of 0.6 where the temperature dependences of the relevant superconducting

parameters are relatively weak. This temperature region

is readily accessible using compact, energy-efficient

closed cycle refrigeration systems.

A review of recent progress in the fabrication of

various types of refractory high temperature Josephson

devices is presented. Certain configurations require

dimensional resolutions approaching 0 . 1 micrometer or less for successful operations. In other structures,

there are problem areas in the growth of suitable thin

films of the superconducting materials of interest and

of suitable barrier layers and controlling the properties of the interfacial regions between these layers and

the superconducting electrodes. Current progress and

accomplishments are summarized and critical areas for

future R&D are identified. Based on recent progress,

the prospects for preparing all-refractory Josephson

devices with onset temperatures exceeding 15K appears

very promising.

Key words: Superconducting devices; Josephson effects;

proximity effect devices; tunnel devices; weak links;

closed cycle refrigeration systems.

1. Introduction

The use of superconductive sensors and circuits as well as cooled circuits employing conventional technologies is slowly gaining acceptance. The

recent development of small, reliable and energy efficient cryogenic refrigeration systems operating in the 10-30K region has satisfied the requirements for such electronic circuits as IR detectors and for cryogenic vacuum

pumps ( "cryopumps" ) . Such cryocoolers are not particularly suited for

superconductive applications as the current generation of superconductive

electronic devices are fabricated using niobium (T ~9.2K) or lead and lead

alloys (T ~7.2K) and an environment in the vicinity of 4.2K must be provided. Tnese low temperatures require the use of an additional low temperatjire stage for the cryocooler such as a Joule-Thomson stage or a wet expansion engine. These approaches introduce additional complexity, weight, cost

195
and an associated reduction in reliability and convenience of use. Nevertheless, a number of industrial and government laboratories have undertaken

programs to develop small, energy-efficient cryocoolers whose characteristics are designed to optimally match the requirements of present generation

superconductive sensors and circuits.

An alternative approach to the problem of integrating superconductive

electronic systems and presently available cryocoolers is the development of

high operating temperature superconductive device technology. If techniques can be perfected for the fabrication of superconductive devices and circuits

using materials with superconductive transition temperatures in excess of

16K, operation in the vicinity of lOK would be practical without imposing

severe restrictions on temperature stability and present day cryogenic

technology might be adequate. In this paper, the current status of high

temperature superconductive device technology will be reviewed, existing

problem areas identified, and recommendations made for possible research and

development

.

2 . Background

A partial list of practical superconductors is given in Table 1 (Devices using superconductors with T below 4.2K are not deemed

"practical"). The present generation of superconductive devices use niobium

or lead alloy technologies and are usually operated in an environment provided by a liquid helium bath. Thus, these devices are operating at reduced

temperatures between 0.5 and 0.6 (see Table 2). At such reduced temperatures, the temperature dependence of material parameters (e.g., superconducting energy gap, critical current, penetration depth, etc.) are relatively weak compared to the behavior close to T . If operation at reduced

temperatures of 0.6 or lower is acceptable, operation in the vicinity of lOK

would be acceptable for devices and circuits fabricated using materials with

T in excess of 16K (see Table II). Thus, we will adopt the definition that

hSgh temperature superconductive devices are those that can be operated at

temperatures in the vicinity of lOK, at a reduced temperature of 0.6 or less

.

The advantages that can be gained by the use of cryocoolers operating

near lOK relative to those that operate in the vicinity of 4.2K is illustrated in Fig. 1. For convenience of comparison, each set of data has been

normalized at lOK. The solid line is the "inverse" Carnot efficiency, that

is, the number of watts of power at room temperature required to produce one watt of refrigeration at the indicated operating temperature. The data

points in Fig. 1 are for characteristics (thermal efficiency, weight, volume

and cost) of crycoolers providing one to several watts of cooling at the

indicated temperature [1], [2]. It is not too meaningful to draw lines

through a given series of data points as it would be too difficult to cor- rect the data for the wide variation in all the other parameters of the

systems. Thus, a shaded region is shown to indicate the trend in these

parameters as a function of operating temperatures. It can be seen that

although 4K operation is theoretically only 2\ times more difficult to

achieve, for practical systems, a factor of about 10 increase is observed in

the parameters such as efficiency, size, weight- and cost. Hence the cryogenic support portions of an integrated superconductive electronic system

can be simplified significantly if lOK operation is feasible compared to 4K

operation.

196
Table 1. Properties of Superconducting Materials of

Possible Technological Importance

Material ^c (K)^ (^)

Mechanical

Properties

Substrate

Temperature

jLeaa/JLieaa A-Lioys 7. 2

Q A n oor t ~K1

Niobium 9. 2 400 Hard RT-*400°C

Mo-Re 15 50 Hard LN^^RT

NbN 16 50 Hard <300°C

NbCN 18 40 Hard ~500°C

V_Si

NB_Sn

17. 3 30 Hard ~800°C

18 3 30 Hard ~800°C

Nb^Ge 23 30 Hard ~900°C

Table 2 . Operating Temperature for Superconducting Devices

Lead and

Lead Alloys Niobium NbN Nb^Ge

Transition Temperature 7.2 K

Usual Operating Temperature 4.2K

Reduced Temperature 0.58

Projected Operating Temperature

9.2K

4.2K

0.48

16K

<0.6

<9.6K

23K

<0.6

<13.8

TEMPERATURE IKI

Fig. 1. Variation of selected parameters of closed cycle cryogenic refrigeration systems as a function of operating temperature. The solid line is

the inverse Carnot efficiency as function of operating temperature. Each

set of data has been normalized to unity at lOK (see text).

197
(

f

1

3. High T Materials I i ^ : ! There is a large volume of literature on the preparation of thin super- j

( conducting films with transition temperatures in excess of 15K [3], [4]. i ! However, for the preparation of a given type of device there may be special ! ' requirements on certain fabrication parameters such as substrate tempera- j

( ture, film thickness, crystalline grain size, nature and degree of granular- i ity or "inhomogeneity, " etc., which may warrant additional activity in I material research. Some of these aspects of thin film preparation will be | identified in the following sections dealing with specific Josephson device j

i configurations. j

I

• 1

4. Proximity Effect Devices i A proximity effect Josephson device can be fabricated by locally suppressing the T of the host film. The initial demonstration of this concept j was by Notarys and Mercereau [5] who used a very narrow strip of normal i material across a film either as an overlay or an underlay to reduce the T i by the proximity effect between a normal metal and a superconductor. Prox- \ imity devices have also been fabricated by reducing the thickness of the I best film over a local region by anodization [5], radiation damage [6], ion | implantation [7], [8], etc. Josephson proximity effect behavior can be I observed whenever the length of the region of suppressed T (that is, the 1 dimension along the direction of current flow) is comparable S:o a character- | istic length which is a function of the normal state parameters. When the

temperature is maintained below the T of the contacts (or "banks") but

above the T' of the localized region, iSiere will be a critical current with } a temperature dependence that will depend exponentially on the ratio of the I characteristic length relative to the physical length of the bridge and on j the difference in temperature between the operating temperature and some ; characteristic temperature which will be close to the suppressed T' of the i region. Josephson effects, both dc and ac, can be observed in this regime | which typically spans a temperature interval of 0.2 to 0.5K. 1 Proximity effect devices have been fabricated from high T supercon- i ducting films using many of the techniques described above. 5he highest j temperature of operation is about 17K reported by Palmer, Notarys and

Mercereau [9] for Nb^ Sn films in which the T of a 1-2 pm long region had I been suppressed by about 0.5K from the original T ~17.5K by selective i anodization. DC SQUIDS made in this fashion showed clear modulation pat- i terns with good signal-to-noise ratios although no minimum detectable field j' change was reported. The devices operated over a temperature range of a few |, tenths of a degree and were reliable, did not degrade under repeated thermal j cycling, etc., and exhibited the other desirable characteristics expected | for devices fabricated from refractory materials. | The proximity effect behavior can be easily observed provided that the i characteristic length associated with the proximity region can be obtained , with the resolution available from existing fabrication techniques. The

primary disadvantage is that the critical current varies exponentially with ' temperature and that the temperature range of operation in the proximity

j mode is quite narrow requiring precise control and regulation of the cryo- i genie environment. !

198 i'
5 . Tunnel Device

The "classical" Josephson device configuration is the oxide tunnel

device. This configuration has been extensively studied theoretically and

extremely good agreement has been obtained between theory and observed

behavior of carefully prepared devices. However, there are serious problems

associated with the fabrication of refractory high T tunnel devices,

j First, it is difficult to grow native oxides with controllable and reproducI ible characteristics on refractory metals, and especially on alloy and comj pounds of these metals [4] . Secondly, almost all of the high T alloys and

compounds require elevated substrate temperatures (300°C and above) for

proper film growth. During the deposition of the top electrode, such elevated temperatures may degrade the properties of the base electrode and may

destroy the very thin (~20a) oxide barrier region.

1 Techniques have been developed for reliably growing oxide barriers on

j the soft superconductors which have very reproducible thicknesses and

I exhibit very small values of excess current below the energy gap voltage in

I the I-V tunneling curve [10]. When these techniques have been applied to

I niobium and other high T materials, noticeably poorer quality tunnel junctions have been obtained ill] , [12]. This is partially due to the fact that

niobium and vanadium, the other refractory material commonly used for high

T alloy preparation, have a great affinity for oxygen and hence the surface

of the base electrode must be thoroughly cleaned by some technique such as sputter etching or glow discharge before the oxide growth is initiated [13].

These cleaning procedures may mechanically damage a shallow region just

below the surface, thus degrading the superconductivity at the surface of

the base electrode. Since the coherence length of these high T materials

is less than lOOA (see Table 1), even a very thin surface layer of damage

will be evident in the properties of the resultant tunnel device (In the case of the soft, low T materials, such as lead and lead alloys, a thin

region of damage at the *^surface of the base electrode does not noticeably

affect device operation as the coherence length in these materials is of

the order of 1,000A or greater). Furthermore, in the case of niobium, there

are a number of natural oxides, sub-oxides etc., some of which are insulators and some conductors and even superconductors (bulk NbO has T ~1.5K)

[13]. Hence depending on the details of the growth process, the oxidi layer

grown on niobium and niobium alloys will have varying electrical characteristics .

In the case of the elemental superconductor, niobium, the group at IBM

has developed techniques for sputter cleaning the surface of the niobium

base electrode, growing an oxide barrier by the Greiner technique and then

sputter treating the surface of the oxide layer to minimize the possibility

of interaction between surface atoms of the oxide layer and the impinging

niobium atoms that form the counter electrode [13]. With these techniques

IBM has succeeded in producing fairly good Nb-Nb^Oc-Nb tunnel junctions as gauged by the low value of the excess current judx oelow the voltage corre- sponding to the energy gap. However, the characteristics of these devices

are still not as good as those obtained for lead alloys devices or for

Nb-Nb20^-Pb alloy devices.

In the case of high T alloy systems, standard oxidation techniques do

not work. When they have Seen used, the quality of the device characteristics are quite poor (see upper trace in Fig. 2) [14]. Noticeably improved

tunneling characteristics have been obtained if the high T base electrode

is first coated with a thin layer (~20A) of some material which readily

oxidizes such as aluminum or silicon and this artifical layer is then

allowed to oxidize [14], [15], [16], [17], [18], [19]. The improvement in

the tunneling characteristics of devices with artificial oxide barriers as compared to devices with native oxide barriers is obvious by comparison of

the several traces in Fig. 2. The concept of using artificial oxide bar199
riers for refractory Josephson devices was apparently first used by

Laibowitz and Mayadas [15] for niobium devices and has subsequently been

adapted by other groups interested in tunneling into high T materials.

Using these artificial oxide barriers and soft low T lead or lead alloy i counter electrodes, tunneling characteristics of reasonable quality have I been obtained. The detailed mechanism of formation of these artificial bar- | riers is not fully understood [20] as a 20A thick film is obviously not con- tinuous and therefore the nature of the resulting oxide barrier probably is

j not continuous and "pinhole free." However, once the optimum thickness for l the oxidized aluminum or silicon layer has been determined, good quality

tunnel junctions with reproducible and controllable thicknesses can be

obtained.

VOLTAGE (mV) Fig. 2. Tunneling characteristics of V_Si-Pb junctions for several types of

tunneling barriers, natural as well as Oxidized silicon layers (see Ref. 14)

Good quality tunneling characteristics have been obtained using artificial oxide barriers on high T base electrodes v/here the counter electrode has

been a soft material (such §s lead or lead alloys), which can be deposited

at or below room temperature [14], [17], [18], [19]. The tunneling characteristics are noticeably poorer if high T materials are used for the

counter electrode- Good high T behavior can only be obtained for the

potentially high T materials, iF they are deposited at elevated temperatures - at least 3uOC and, in certain cases, at temperatures as high as lOOOC. These elevated temperatures can have disastrous effects on the

properties of the base electrode and the oxide barrier region that has been

grown on the base electrode. It is well known that oxygen can diffuse

readily through niobium and niobium alloys at relatively low temperatures.

Diffusion of oxygen can either suppress or destroy superconductivity in the

layer into which the oxygen has diffused. If superconductivity is suppressed at the surface of the electrode, the desired abrupt transition

between uniform superconducting and insulating regions will be smeared out

and the quality of the tunneling characteristic will be degraded, and excess current will be observed below the energy gap voltage. Hence, the deposition of the high T counter electrode for an all-refractory superconducting

tunnel device presents a formidable problem.

200
ii

ll

1 An attempt to fabricate an all high T Josephson device has been

I reported recently by Tarutani et al . [21], who used Mo-Re as the counter

I electrode material. The system of Mo-Re, which in the bulk can have a T as

I high as 12. 2K, can be deposited at temperatures near or below room temperaI ture and, if the deposition conditions are suitable, the compound Mo_Re„

will be formed with a T as high as 11. 2K. They report on the preparation

I of tunnel devices with V^Si as the base electrode (T ~16.4K when deposited

at 900°C), an artificial oxidized silicon barrier (native oxide barriers did

j not exhibit tunneling) and Mo-Re counterelectrodes . Films of Mo-Re deposited at ambient temperatures did not exhibit superconductivity as interdiffusion of atoms between the Mo-Re, as it was being deposited, and the

I remainder of the structure probably occurred as the temperature of the

device was raised to about 130C during the Mo-Re deposition. When the deposition was made at LN„ temperatures, the device probably remained below lOOK

and Mo_Re2 was formea which, even after being warmed to room temperature,

j exhibited superconductivity with a broad transition extending from 10. 5K

down to 8.3K. The article did not contain details about the tunneling characteristics of the devices nor did it indicate the highest temperature at

I which Josephson effects could be detected.

Another likely candidate for the counter electrode in an all refractory

tunnel device is NbN. This material is normally deposited at temperatures

in the range of 300-500C and transition temperatures of the order of 15K are obtained for films of thickness of 1000 S or greater. If the device is

reactively sputtered at ambient temperature, T is reduced to the vicinity

of 10-12K. An attempt to fabricate an all ifibN tunnel junction has been

reported by Shinoki et al. [22], who used an artificial layer of oxidized

silicon to form the tunneling barrier. The base electrode was deposited at

500C and exhibited T of about 15K. In order to minimize degradation of the

base electrode and S:he oxidized silicon barrier region, the NbN counter

electrode was deposited at near ambient temperatures. The tunneling characteristics at 4.2K show a distinct energy gap at 4mv which suggests that the

energy gap for the upper film is below lOK. Although no specific value is

quoted, the excess current below the gap appears to be large and the device

as reported was not suitable for switching application.

6 . Semiconductor Barrier Devices

Another Josephson tunneling configuration employs a semiconductor barrier several hundred angstroms thick instead of the oxide barrier which is

normally tens of angstroms thick. The Naval Research Laboratory has a joint

effort with Sperry Research Center to fabricate all refractory superconducting tunnel devices with CVD deposited polysilicon barrier [23], [24]. The

base NbN electrode is reactively sputter-deposited at a substrate temperature of about 500C which results in a T of about 15K. A several hundred

angstrom thick silicon layer is grown bySzVD on the base electrode which is

held at a temperature of 300-500C and a counter electrode of NbN is then

reactively sputtered at (nominally) ambient temperatures. After fabrication, the base electrode has a T -13. 5K while the T of the counter electrode is found to be about 12K. ^Josephson currents were observed at temperatures slightly below lOK. Unfortunately, the preliminary results were

such that it could not be unequivocally established whether the supercurrent

was due to tunneling through the silicon barrier or along shorts across the

barrier region. At the time of the preparation of this communication,

Sperry had improved their techniques for depositing silicon layers on NbN

and new all refractory, semiconductor barrier devices will be fabricated in

the near future.

201
7 . Weak Link/Constriction Devices

Another class of device configuration which can exhibit Josephson \ effects is a structure which contains a very severe reduction in width over i a very short distance - the so-called "weak link". The theoretical con- j straint for the observation of Josephson effect is that length (and width) ' of the constriction must be comparable to or smaller than three times the ! superconducting coherence length of the material [25]. In the case of the , mechanically soft, low T materials, such as tin and aluminum, coherence 1 lengths are of the order oT thousands of angstroms and ideal weak links have : been fabricated by conventional optical lithographic techniques or scribing j with sharp diamond points. However, the coherence length of materials tends j to decrease with increasing transition temperature and materials such as 1 niobium and the high T materials have coherence lengths so small that con- ! ventional techniques cannot be used to fabricate constrictions that are ! small enough to satisfy the conditions for obtaining ideal Josephson I behavior. I In the case of niobium, which has a coherence length of the order of

400A, recently developed "contamination resist" techniques have been used to

fabricate constrictions with widths of the order of 50o8 and lengths about

1200a which have shown ideal Josephson behavior both in their response to ] incident microwave radiation [26] and when used to fabricate SQUID magneto- | meters [27]. The techniques for fabricating these so-called "nanobridges" j (so named as they are smaller than "microbridges" ) cannot be employed to ! fabricate Josephson weak levels in the refractroy alloy materials, espe- [ cially those that have transition temperatures in excess of lOK. These

materials have coherence lengths of the order of 5o8 or less (see Table I),

j a dimension not reachable with any current device fabrication technique. | There are a number of references where the authors have reported ; phenomena at temperatures approaching 17K which they have attributed to the \ Josephson effect. Golavashkin et al. [28], [29], has reported on a Nb_Sn \ device in which they mechanically scribed two constrictions 1-5 |jm wide and j 1-2 pm long to produce what is supposedly a DC SQUID. The device was sen-

' sitive to changes in the applied magnetic field in a manner resembling a DC

j SQUID although the response behavior was poor compared to that observed for | DC SQUIDS fabricated from low T materials. The poor quality of the i response might be due to operation^by some type of synchronized "flux flow"

which one would expect to observe for the case of apparent bridge dimensions i at least 100 times larger than the coherence length of Nb„Sn. Another pos- j sible explanation for this type of response (from a device with constrictions ; with these dimensions) is that there are flaws (such as cracks or grain

boundaries) within the constricted region such that the dimensions at some point are very small compared to the apparent dimensions. Another possi- I bility is that the region of the film where that constriction was produced

was not uniform and that the operation of the device is attributable to ; granular or inhomogeneous superconductivity (see below). Until the true i nature of the constriction can be established, reservations must be exer- cised before attributing the observed behavior to true Josephson effect. j! The same reservations should be made in considering the recent report j by Lee and Falco [30] on the observation at temperatures near 17K of RF | induced microwave steps in the I-V characteristics of constrictions made in i Nb^Sn films. A constriction with dimensions of the order of micrometers was sculptured into the film and then the transition temperature of the con- i striction was adjusted downward by applying across the constriction voltage ^ pulses with microsecond duration. By a mechanism as yet not understood,

this technique permitted a systematic and controlled depression of the T of | the constriction. Using a device in which the T had been depressea to ' about 17K from an original value of about 18K, t'hey observed RF induced ! steps in the device I-V curve at voltages corresponding to the integer har202
monies of the Josephson frequency: no subharmonic or non-integer harmonic

steps were obtained. This type of response is expected for ideal Josephson

behavior while synchronized flux flow, which would be expected for a con- striction in Nb^Sn with these dimensions, would also have exhibited subharmonic or non-integer harmonic steps. The modification of the film that

occurred during the "pulse" treatment is not understood: small fractures

may be created or, possibly, local heating may expedite the diffusion of

oxygen along grain boundaries thus creating a region within the constriction

which exhibits inhomogeneous superconductivity. (In the latter case, as will be discussed below, the coherence length does not impose a severe con- straint on the dimensions of an inhomogeneous weak link. ) Notwithstanding

the lack of knowledge about the details of the physical and electrical

nature of the constriction after the pulse treatment, the preliminary

results appear quite convincing that true Josephson effects have been

observed in refractory high T materials at temperatures as high as 17K.

Furthermore, this fabrication ^technique should be suitable for producing

Josephson devices in other refractory materials. This technique, which was first employed in producing superconducting weak link constrictions in

niobium films by Duret et al. [31], might be considered to be the thin film

equivalent of the bulk mechanical point contact configuration. In both

cases, the device is adjusted while observing the superconducting or electrical properties, the thin film version by repeated pulsing while the bulk

contact by very careful adjustment of the tension on the point. (Unfortunately the thin film version is not reversible while the point contact is.)

The pulsing technique may permit extensive investigations of Josephson

effects in high T^ films, much the same way that the point contact was an extremely valuable tool in studying bulk materials. Although this technique

may prove to be very useful in making single Josephson devices, it is doubtful that it can be employed to fabricate complex circuits and/or arrays of

Josephson devices.

8 . Granular Weak Links

As pointed out in the previous section, theory predicts that ideal

Josephson behavior should be observed in weak link constrictions only if

their dimensions are of the order of or less than three times the coherence

length of the material [25]. For refractory alloy materials, coherence

lengths are of the order of 100 A or less and thus the observation of

Josephson behavior requires constrictions with dimensions of the order of

300 A or less, a value that cannot be achieved by conventional electron beam

lithography and is at the limit of what may be achieved with "contamination

resist" lithography developed by IBM.

However, the group at NRL has shown that weak link constrictions in

refractory alloy films can exhibit Josephson effects with micrometer sized

dimensions if the weak link region is sufficiently thin to exhibit granular

or inhomogeneous superconducting behavior [32]. When a film of, for

example, niobium or niobium nitride of thickness of lOOGA is thinned down by

anodization or ion milling, the transition temperature begins to decrease

(see Fig. 3). The variation of the transition temperature with film thickness obeys the following relation

T^(d) = T^ (00) exp[ - dyd ] For both niobium and niobium nitride, the constant d was found to be about

lOA. The significance of this parameter is that T ^f the film appears to

be reduced by the proximity effect with non-superconducting layers of 5A on each surface of the film.

As the film thickness is reduced, a critical value is reached, about

lOOA for niobium nitride and about 5oK for niobium films, below which there

203
IS a noticeable decrease in the critical current density of several orders of

magnitude (see Fig. 3). The same reduced value of critical current density

values is observed for thinner films. However, the T varies smoothly

through this critical thickness value and obeys the above^ relation for all

values of film thickness.

10 50 100 600 1000 3000

FILM THICKNESS (Angstroms) Fig. 3. Variation of the superconducting transition tempeature and critical current density of niobium nitride films as a function of film thickness

The decrease in criti'cal current density is believed to be due to a transition from bulk conduction for the thicker films to a granular or hopping conduction for films with thicknesses less than the critical value.

As the film is deposited, it initially grows as isolated islands of materials which develop to sizes of the order of lOOA. As the second layer of

grains begin to grow, they touch and "bulk-like" conduction occurs. Conversely, when a thick film is thinned until the thickness of the region

becomes comparable to the grain size, the conduction changes from bulk-like

to tunneling between the isolated grains. The proposed nature of the con- duction mechanism for thin films is supported by the observation that many

of these films exhibit a negative coefficient of resistance for temperatures

above T^ which can be attributed to hopping conduction between isolated

grains. Similar behavior has been observed in granular aluminum and tin

films which were deposited in an oxygen partial pressure which aided in the

formation of isolated grains coated with oxide layers.

SQUIDS have been fabricated using granular regions with dimensions of

the order of one micrometer in length which exhibited good Josephson

behavior [33], [34]. (Granular regions with lengths greater than several

micrometers exhibited rather poor or smeared out Josephson responses.)

Depending on the degree of granularity of the weak link region, the width of

the localized region can be varied from submicrometer for films which were marginally granular to widths greater than 40 micrometers for weak links

which were clearly granular in nature. For niobium nitride SQUIDs with weak

links that were clearly in the granular regime, Josephson behavior could be

followed from T , which for some devices was as high as lOK, down to about

1.2K, the loweit temperature that could be reached (see Fig. 4). Ideal

Josephson behavior was obtained for critical currents greater than (!)q/L (the

condition for hysteretic behavior required by the Kurkijarvi theory) but

less than about 6 /L [34], [35]. The degradation of response for large

values of critical ^current is not understood but is being investigated.

Several of these devices have been examined at microwave frequencies (about

10 GHz) and exhibited Josephson behavior with extremely good signal-to-noise

ratios, over a corresponding wide temperature range from just below T down

to about 1.2K [34]

.

204
Iff (ARBITRARY) Fig. 4. I-V characteristics of NbN RF SQUID containing a granular weak link

measured at 20 MHz as a function of temperature. The onset temperature for

observing Josephson behavior was lOK while 1.27K was the lowest temperature

that could be achieved in the particular apparatus (see Refs. 34 and 35)

If the weak link region has been thinned down so that it is just mar- ginally granular in nature, Josephson behavior can be observed at much

higher temperatures but over a restricted temperature region. For example,

the data in Fig. 5 is the response of a NbN SQUID containing a marginally

granular weak link which exhibits an onset of SQUID behavior at near 14K.

The response can be followed down to about 13K before the critical current

becomes too large. As described in the previous paragraph, ideal magnetic

flux noise, characteristic of the temperature, is observed for critical cur- rents greater than 1 (t) /L but less than about 6 (t) /L, corresponding to a one degree temperature range of operation. The quality of the response and the

temperature range over which it is observed is superior to the previously

reported response for any RF-biased SQUID fabricated from refractory alloy

materials

.

The physics behind the observed operation of these granular weak link

constriction devices is not understood [32], [35]. The grain size in these

films is about 10(^ and thus in a 1 micrometer by 1 micrometer region, there

are more than 10 grains. However, a granular region of this dimension

built into an ^F biased SQUID behaves as if it were a single ideal Josephson

device--the 10 grains are strongly coupled together and appear to be characterized by a single quantum mechanical phase difference common to the

entire array of grains.

205
Thus far, granular behavior has been observed in two refractory thin

film systems - niobium and niobium nitride. If similar type of behavior can be observed in higher T material systems, it may be feasible to fabricate

weak link structures with micrometer dimensions which may show onsets of

SQUID behavior at temperature near 15K and thus, could be operated near lOK

(that is, at a reduced temperature of 0.6) where the temperature coefficients of the various superconducting parameters are weak. At these operating temperatures, the cryogenic environment could be provided by small, compact, energy-efficient cryogenic refrigeration systems.

Materials . There are a fairly large number of superconducting materials that are potentially suitable for the fabrication of high temperature

devices and circuits. Many of these materials have been prepared in thick

film form (thicknesses greater than 1 micrometer) with the purpose of

exploring their use in high power superconducting applications. However,

there has been only limited work on preparing thin refractory films (thick- nesses of the order of or less than 1,000A) in a form that would be suitable

for fabricating Josephson devices and superconducting electronic circuits.

Additional effort will be required to develop procedures for preparing fine

grain structures, that adhere well to substrate materials and that can be

prepared with desired properties under fabrication conditions that are com- patible with multi-layered device processing.

Device Processing . In the fabrication of tunnel junctions using all

refractory high temperature superconducting electrodes, the major problem

areas are the barrier region, either the natural oxide or artificial barriers, suitable masking, and the deposition of the refractory counter electrode without degrading the base electrode or the barrier region (see

Table 3). The fabrication of weak link constrictions in refractory materials that can exhibit ideal Josephson behavior requires device geometry reso- lutions in the 300A regime, which is beyond the capabilities of presently

available device processing technology. Granular weak links have been

fabricated with micrometer dimensions with operating temperatures as high as 14K. Additional materials research would be required for the fabrication of

granular systems which have transition temperatures at values in excess of

15K so that devices made from these materials can be operated at lOK at a reduced temperature of 0.6.

9. Conclusions

Table 3. Problem Areas for Fabricating High Temperature

Josephson Devices (T»p > lOK)

Type of Structure

Highest Reported

Operation Problem Area

Tunnel Junction Barrier layer growth

Defining masks

Counter electrode

Proximity Effect

Device

17K Limited temperature range

Exponential I dependence

Weak Link/Constriction Small dimensions (~10oS)

Granular/Inhomogeneous

Weak Link

13K Preparation of granular

specimens of high T^

materials

206
Prospects for Future Progress . Despite the limited amount of research

activity directed at' the fabrication of all refractory high operating temperature Josephson devices and circuits, the prospects are quite good that

these materials will be used in the near future in Josephson technology.

The primary obstacles to their use are materials problems such as producing

controlled barrier layers and suitable counter electrodes in the case of

tunnel junctions, and producing high T materials which exhibit granular

behavior in the case of weak links. These are quite serious obstacles,

especially given the refractory nature of the materials and the affinity of

these materials for such undesirable impurities as oxygen. However, the

modest efforts to date at IBM, NRL and in Japan have yielded quite encouraging results.

All refractory devices would be extremely rugged and reliable and would

withstand repeated thermal cyclings to cryogenic operating temperatures

.

These properties are not characteristic of the presently used lead alloy

Josephson device technology but are essential if Josephson technology is to

become a viable one, either on a commercial basis or for military applications. In addition, these refractory materials often have superconducting

transition temperatures in excess of 15K which makes possible their operation near lOK (at a reduced temperature of 0.6). This would permit the use of

fairly compact, highly reliable closed cycle refrigeration systems for

deployment with Josephson sensor systems and, possibly small Josephson circuits and systems. It would appear that large scale digital processing systems would continue to be operated at liquid helium temperatures as the

liquid is probably essential for adequate cooling of the very complex packaging associated with such systems.

The ruggedness and reliability of Josephson devices and circuits fabricated from refractory materials provide the impetus that will motivate the

developing of this technology. Current activities are limited but should

expand in the near future so that one may confidently expect to see allrefractory Josephson devices within the next five years and fairly complex

integrated circuit structures fabricated with refractory metallization

within the next ten years. The advent of such circuit technology is crucial

to the development of a viable superconducting electronic technology and

their development will be pursued despite the serious materials problems

that must be overcome.

10. References

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Technical Note 655, June 1973 (U.S. Government Printing Office,

Washington, D.C. 20402, catalogue number C13.46:655), unpublished.

[2] Daunt, J. G. , and Goree, W. S., "Miniature Cryogenic Refrigerators,"

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,

unpublished.

[3] Beasley, M. R. , "Improved Materials for Superconducting Electronics,"

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207
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[5] Notarys, H. A. and Mercereau, J. E., "Proximity Effect Bridges and

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Notarys, H. A., "Comparative Studies of Ion-Implant Josephson-Effeet

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Single Crystal Niobium Films," J. Appl. Phys. 41, 2748 (1970). j [12] Broom, R. F., Jaggi, R., Laibowitz, R. B., Mohr, Th. 0. and Walter, W.

,

"Thin Film Josephson Tunnel Junctions with Niobium Electrodes,"

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Physics; Otaniemi, Finland; August 1975 (North Holland Publishing

Co., Amsterdam); Vol 4, p. 172.

[13] Broom, R. F. , Laibowitz, R. B., Mohr, Th. 0. and Walter, W. , "Fabrication and Properties of Niobium Josephson Tunnel Junction," IBM J. Res. Develop. 24, 212 (1980).

[14] Moore, D. F., Zubeck, R. B., Rowell, J. M. and Beasley, M. R. , "Energy Gap

of A-15 Superconductors Nb_Sn, V_Si and Nb„Ge measured by Tunneling," Phys. Rev. B20, 2721 X1979).

[15] Laibowitz, R. B. and Mayadas, A. F. , "Josephson Junctions with Nb/Al Composite Electrodes," Appl. Phys. Lett. 20, 254 (1972).

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and Viggiano, J. M. , "Proximity Effect Tunnel Junctions with Barriers Formed from Amorphous Alloys," Jour, de Phys. 39, C6-1340

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[17] Yeh, J. T. C. and Tsuei, C. C, "Tunneling Studies on Thin Film Nb-Al

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[18] Buitrago, R. H. , Goldman, A. M. , Toth, L. E. and Cantor, R. "Preparation of Nb„Ge Superconducting Tunneling Junctions," IEEE Trans, on

Magnetics, mG-15 , 589 (1979).

[19] Howard, R. E., Rudman, D. A. and Beasley, M. R. , "Josephson Properties

of Nb^Sn/Pb tunnel Junctions," Appl. Phys. Lett. 33, 671 (1978).

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208
[22] Shinoki, F. , Takada, S., Kosada, S. and Hayakawa, H., "Fabrication of

High Quality NbN/Pb Josephson Junctions," Jap. Jour. Appl. Phys.

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[23] Kroger, H. , Potter, C. N. and Jillie, D. W. "Niobium Josephson Junctions with Doped Amorphous Silicon Barriers," IEEE Trans, on

Magnetics, MAG-15 , 488 (1979).

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[25] See for example, Likharev, K. K. , "Superconducting Weak Links," Rev. Mod.

Phys. 51, 101 (1979).

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,

"Josephson Effect in Nb Nanobridges , " Appl. Phys. Lett. 35, 891

(1979)

.

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SQUID," Appl. Phys. Lett. 37, 656 (1980).

[28] Golovashkin, A.I., Levchenko, I.S., Lykov, A.N. and Makhashvili, L.I.,

"Quantum Interference on Nb_Sn Thin Film Contacts at Hydrogen Temperatures," JETP Lett. 24, 521 (1976).

[29] Golovashkin, A. I., Levchenko, I. S. and Lykov, A. N. , "Quantum Interference at Nb and Nb_Sn Superconducting Microbridges, " Soviet

Physics Solid State 18, 2121 (1977).

[30] Lee, T. W. and Falco, C. M. , "Josephson Effects in Nb^Sn Microbridges,"

Applied Superconductivity Conference, Santa Fe, N.M., October

1980; IEEE Trans, on Magnetics, Spring 1981, to be published.

[31] Duret, D., Bernard, P. and Zenatti, D., "A UHF Superconducting Magnetometer Utilizing a New Thin Film Sensor," Rev. Sci. Instr. 46, 475

(1975)

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[32] Claassen, J. H., Cukauskas, E. J. and Nisenoff, M. , "Granular We^k Link

Josephson Devices," Inhomogeneous Superconductors-1979, ed.

Gubser, D. U. , Francavilla, T. L., Leibowitz, J. R. and Wolf, S. A.

(American Institute of Physics, New York, 1980), p.- 169.

[33] Rachford, F. J. and Cukauskas, E. J., "High T Refractory Thin-Film

Microwave SQUIDs," Appl. Phys. Lett. 35, 891 0.979).

[34] Cukauskas, E. J. and Nisenoff, M., "Intrinsic Noise Characteristics of

NbN SQUIDS," J. Appl. Phys. 52 (Feb 1981), accepted for publication.

[35] Cukauskas, E. J. and Nisenoff, M. , "Noise Properties of Thin Film

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209
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October 6 & 7, 1980

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P. 0. Box 218

Yorktown Heights, NY 10598

M. X. Francois

Laboratoire de Thertnodynamlque des Fluldes

Campus Unlversltaire Bat 502 Ter

91405 Orsay

Cedex

France

J. Bindslev Hansen

Physics Lab. - Univ. of Copenhagen

Unl versltetSparken 5 DK-2100

Copenhagen

Denmark

Max G. Gasser

NASA Goddard SFC

Code 713

Greenbelt, MD 20771

Robert G. Hansen

R. G. Hansen & Associates

1324 State Street

Santa Barbara, CA 93101

Peter E. Gifford

President, Cryomech, Inc. 314 Ainsley Drive

Syracuse, NY 13210

W. H. Hartwig

Dept. of Electrical Engineering

P. 0. Box 7728

University of Texas at Austin

Austin, TX 78712

Dave Giguere

Hughes Aircraft Company

EED Bidg. 240

3100 Lomitom Blvd.

Torrance, CA 90509

Wmiam Haskin

Adv. Cryogenic Systems Group

Environmental Control Branch

Dept. of the Air Force

Wright-Patterson Air Force Base, OH 45433

Mike Goldowsky

Philips Lab. 345 Scarborough Road

Briarcliff Manor, NY 10510

C. Heiden

Institut fur Angewandte

University of Giessen

Heinrich-Buff-Ring 16 D-6300 GieBen

W. Germany

Geoffrey Green

DTNSRDC (Annapolis Lab.)

Code 2712

Annapolis, MD 21402

Thomas E. Hoffman

Arthur D. Little, Inc. Acorn Park

Cambridge, MA 02140

212
ATTENDANCE LIST, continued

R. Hollman

Physics Department

Stanford University

Stanford, CA 94305

Kenjiro Kasai

Kasado Research Dept.

Hitachi Ltd.

794 Kudamatsu-City

Yamaguchi-Ken 744

Japan

Stuart Horn

U. S. Army Night Vision Lab. NV & E.O. Lab.

Ft. Bel voir, VA 22060

Dale Kennedy

AFWAL/FIEE

Wright- Patterson AFB Ohio 45433

Yoshihiro Ishizaki

University of Tokyo

Tokyo, Japan

Peter J. Kerney

CTI Cryogenics

Kelvin Park

Waltham, MA 02254

Lyie M. Ishol

Code 792

Naval Coastal Systems Center

Panama City, FL 32407

A. L. Johnson

Aerospace Corp.

D-5 P. 0. Box 92957

Los Angeles, CA 90009

Dean L. Johnson

Jet Propulsion Laboratory

M. S. 238-737

4800 Oak Grove Drive

Pasadena, CA 91106

T. Kondo

Engineering Research Institute

Faculty of Engineering

University of Tokyo

Bunkyo-Ku, Tokyo

Japan

Tadashi Kondoh

Manager, 2nd Technical Lab.

Aisin Seiki Co. , Ltd.

1-Asahi-Machi , 2-Chome

Kariya City, Aichi Pref.

448

Japan

Herbert Korf

AGE-Telefunken

Theresienstr. 2 D-7100 Heilbronn

W. Germany

Yasuharu Kami oka

University of California

7671 Boelter Hall

Los Angeles, CA 90024

Tomoo Kurumada

Osaka Oxygen Industries, Ltd. Utajima, Nishiyodogawa-Ku

Osaka

Japan

213
ATTENDANCE LIST, continued

Kurt Kwasnitza Del Linenberger

ETH - Zurich NBS-Thermophysical

Institute of Solid State Physics Properties Division

Zurich Boulder, CO 80303

Switzerland

Rudy Latasa William A. Little

National Radio Astronomy Observatory Physics Department

P. 0. Box 2 Stanford University

Green Bank, WV 24944 Stanford, CA 94305

William Lawless

Lake Shore Cryotronics

64 E. Walnut Street

Westerville, OH 43081

Ralph C. Longsworth

Air Products & Chemicals, Inc. Advanced Products Department

Box 2802

Allentown, PA 18105

Otto Ledford

Aerospace Corp. Box 92957

Los Angeles, CA 90009

J. C. Lottin

Commissariat Al'Energie Atomique

C.E.N. Saclay-Orme des Merisiers, Stipe

Postale No. 2, 91-Gif-Sur-Yvette

France

Daniel Lehrfeld

Magnavox

34 Greenwood Drive

New City, NY 10956

K. Lueders

Freie Univ. Berlin

Each Physik (FB 20)

Institut fur Atom & Feskorperphysi

k

Koenigin-Luise-Str. 28/30

1000 Berlin 33

W. Germany

Robert Lepobsky

CTI, Cryogenics

266 Second Ave. Waltham, MA 02154

John T. Lynch MIT - Lincoln Labs.

Room 329-Box 73

Lexington, MA 02173

James Lester

Ball Aerospace

Box 1062

Boulder, CO 80306

Thomas J. Marusak

Mechanical Technology Incorporated

968 Albany- Shaker Road

Latham, NY 02110

214
ATTENDANCE LIST, continued

Yoichi Matsubara

Atomic Energy Research Institute

N1hon University

Kanda-Surugadal Chiyoda Ku Tokyo

Japan 101

George Patton

DTNSRDC (Annapolis Lab.)

Code 2712

Annapolis, MD 21402

Ken Myrtle

Dept. of Physics

Simon Fraser University

Burnaby B. C. V5A156

Canada

Ousan Petrac

Jet Propulsion Lab, CIT

4800 Oak Grove Drive, 183/401

Pasadena, CA 91103

Lawrence G. Nelsen

Rockwell International

10202 Flintndge

Villa Park, CA 92667

Warren Pierce

Lake Shore Cryotronics

64 E. Walnut Street

Westerville, OH 43081

M. Nisenoff

Naval Research Laboratory

Code 6854

Washington, D. C. 20375

H. Quack

Sulzer Brothers Limited

CH-8401 Winterthur

Switzerland

J. J. Orth

Hughes Aircraft Company

Culver City, CA 90230

Ray Radebaugh

NBS - Thermophysical

Properties Division

Boulder, CO 80303

Isao Oshima

Osaka Oxygen Industries, Ltd. Utajima, Nishiyodogawa-Ku

Osaka

Japan

Richard M. Rail

Hughes Aircraft Company

EED-Bldg. 240/1 13 3100 W. Lomita Blvd.

Torrance, CA 90509

J. .A. Pals

Philips Research Lab. Eindhoven

Nederland

Piter Roos Magnavox

46 Industrial Ave. Mahwah, NJ 07430

215
ATTENDANCE LIST, continued

Pierre M. Roubeau Allan Sherman

Commissariat A TEnergie Atomique NASA Goddard Space Flight Cente

C. E. N. Saclay -Orme des Merisiers, Boite Code 713

Postale No. 2, 91-Gif-Sur-Yvette Greenbelt, MD 20771

France

S.C. Russo

Bidg. 12/V137

Hughes Aircraft Company

Culver City, CA 90230

Arnold H. Silver

Aerospace Corp.

Box 92957

Los Angeles, CA 90009

Ronald E. Sager

S.H.E. Corporation

4174 Sorrento Valley Blvd.

San Diego, CA 92121

J. F. Skinner

Hughes Aircraft Company

Bldg. 5 M.S. B151

Culver City, CA 90230

Raymond Sarwinski Hans A. Steinherz

S.H.E. Corporation CTI Cryogenics

4174 Sorrento VAlley Blvd. Kelvin Park

San Diego, CA 92121 Waltham, MA 02254

E. Saur

Institut fur Angewandte Physik

der Universitat Giessen

Heinrich Buff-Ring 16

6300 GieBen

West Germany

William Steyert

Los Alamos Scientific Lab. Mail Stop 764 - P. 0. Box 1663

Los Alamos, NM 87545

Aron Sereny

Philips Lab. 345 Scarborough Rd. Briarcliff Manor, NY 10510

Philip Studer

NASA Goddard SFC

Code 716

Greenbelt, MD 20771

G. W. Shepherd

Magnavox

46 Industrial Ave. Mahwah, NJ 07430

Donald B. Sullivan

Division 724.03 - Room 2137

National Bureau of Standards

325 Broadway

Boulder, CO 80303

216
ATTENDANCE LIST, continued

Jean-Jacques Thibault

L'Air Llquide, Centre

Claude-Delorme, Boite

38360 Sassenge

FRANCE

d' Etudes Cryogeniques

Postal e No. 15

Lawrence A. Wade

Lockheed Research Lab.

3251 Hanover Street

Bldg. 205 - Dept. 52-32

Palo Alto, CA 94304

K. D. Hmmerhaus

Engineering Center ADl-25

University of Colorado

M. S. 423

Boulder, CO 80309

G. Walker

Dept. of Mechanical Engineering

2920 24 Ave. N.W. University of Calgary

Calgary, Alberta

CANADA

Bryan C. Troutman

IBM 2L50

18100 Frederick Pike

Gaithersburg, MD 20760

Sander Weinreb

National Radio Astronomy Observatory

P. 0. Box 2 Green Bank, WV 24944

E. Tward

Jet Propulsion Laboratory

M. S. 183-901

4800 Oak Grove Drive

Pasadena, CA 91103

Nancy K. Welker

Lab for Physical Sciences

4928 College Avenue

College Park, MD 20740

A. R. Urbach

Ball Aerospace Corp. Box 1062

Boulder, CO 80306

R. White

Adv. Cryogenic Systems Group

Environmental Control Branch

Dept. of the Air Force

Wright-Patterson AFB Ohio 45433

Eugene Urban Calvin Winter

Es. 63, NASA Simon Fraser Univ.

Marshall Space Flight Center Dept. of Physics

Huntsville, AL 35812 Burnaby B. C. V5A156

Canada

R. J. Vincent

ERG, Inc. Lowell and 57th Street

Oakland, CA 94608

James E. Zimmerman

Division 724.03 - Room 2137

National Bureau of Stds.

325 Broadway

Boulder, CO 80303

* U.S. GOVERNMENT PRINTING OFFICE : 1981 0 - SHt-SlB : QL 3

217
NBS-114A IREV. 2-8C) U.S. DEPT. OF COMM. BIBLIOGRAPHIC DATA

SHEET (See instructions)

1. PUBLICATION OR

REPORT NO. NBS SP 607

2. Performing Organ. Report No. 3. Publication Date

May 1981

4. TITLE AND SUBTITLE

Refrigeration for Cryogenic Sensors and Electronic Systems

Proceedings of a Conference held at the National Bureau of Standards,

Boulder, CO, October 6-7, 1980

5. AUTHOR(S)

J. E. Zimmerman, D. B. Sullivan, and S. E. McCarthy, Editors

6. PERFORMING ORGANIZATION (If joint or other than NBS. see /n struct/on s)

NATIONAL BUREAU OF STANDARDS

DEPARTMENT OF COMMERCE

WASHINGTON, D.C. 20234

7. Contract/Grant No. 7240187

8. Type of Report & Period Covered Final

9. SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (Street. City. State, ZIP)

International Institute of Refrigeration - Commission A 1/2

Office of Naval Research - Naval Research Laboratory

Cryogenic Engineering Conference

National Bureau of Standards

10. SUPPLEMENTARY NOTES

Library of Congress Catalog Card Number: 81-600038

Document describes a computer program; SF-185, FlPS Software Summary, is attached.

11. ABSTRACT (A 200-word or less factual summary of most si gnificant information. If document includes a si gnificant

bi bliography or literature survey, mention it here)

This document contains the proceedings of a meeting of refrigeration

specialists held at the National Bureau of Standards, Boulder, CO, on October 6 and 7, 1980. Participation included representatives of industry,

government, and academia. The purpose of the meeting was to discuss progress

in the development of refrigeration systems which have been specialized

for use with cryogenic sensors and electronic systems. The meeting focused

primarily on the temperature range below 20 K and cooling capacity below 10 W. The meeting was jointly sponsored by the International Institute of

Refrigeration-Commission A 1/2, the Office of Naval Research, the Naval

Research Laboratory, the Cryogenic Engineering Conference, and the

National Bureau of Standards.

12. KEY WORDS (Six to twelve entries; alphabetical order; capitalize only proper names; and separate key words by semicolons)

Cryocoolers; cryogenic sensors; helium; refrigeration; superconducting devices.

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