NisrUNITED STATES

DEPARTMENT OF COMMERCE

NATIONAL INSTITUTE OF STANDARDS

AND TECHNOLOGY

NATL INST, OF STAND & TECH R.I.C.

A111D3 MTflOtib

NIST

PUBLiCATiONS NIST Technical Note 1343

Analytical Model for the

Refrigeration Power of the

Orifice Pulse Tube Refrigerator

Peter J. Storch

Ray Radebaugh

James E. Zimmerman

Reservoir

Volume

Piston K/X//X/

100

.U5753

#1343

1990

C.2

Regenerator

Orifice

Heat

Exclianger

Pulse Tube

Heat

Exchanger

mo

Analytical Model for the

Refrigeration Power of the

Orifice Pulse Tube Refrigerator

Peter J. Storch

Ray Radebaugh

James E. Zimmerman

Chemical Engineering Science Division

Center for Chemical Technology

National Measurement Laboratory

National Institute of Standards and Technology

Boulder, Colorado 80303-3328

Sponsored by

NASA/Ames Research Center

Moffett Field, California 94035

\. /

U.S. DEPARTMENT OF COMMERCE, Robert A. Mosbacher, Secretary

NATIONAL INSTITUTE OF STANDARDS AND TECHNOLOGY, John W. Lyons, Director

Issued December 1990
National Institute of Standards and Technology Technical Note 1343

Natl. Inst. Stand. Technol., Tech. Note 1343, 94 pages (Dec. 1990)

CODEN:NTNOEF

U.S. GOVERNMENT PRINTING OFFICE

WASHINGTON: 1990

For sale by the Superintendent of Documents, U.S. Government Printing Office, Washington, DC 20402-9325
NOMENCLATURE

PTR = pulse tube refrigerator

BPTR = basic pulse tube refrigerator

OPTR = orifice pulse tube refrigerator

RPTR = resonant pulse tube refrigerator

a = slope of average temperature profile

Ag = gas cross—sectional area Cp = specific heat of gas

f friction factor

L length of the tube

h specific enthalpy

fl enthalpy flow

<fi> = time—averaged enthalpy flow over one cycle

k thermal conductivity

Mt = total mass of gas in the table

rii = total mass flow rate

rhh = mass flow rate through orifice

rile = mass flow rate at the cold end

nicp = total mass in compressor

m,cp

Pd

mass flow rate from the compressor

total dynamic pressure

PaQc

Qh

Re

time—averaged pressure

heat flow rate into tube at the cold end

heat flow rate out of tube at the hot end

Reynolds Number

R

Th

Td

gas constant in mass units

hydraulic radius

total dynamic temperature

111
T = dead volume temperature (ambient)

% = time—averaged temperature profile in tube

% = cold end temperature (x=0)

% = hot end temperature (x=L, ambient)

%p = compressor temperature (ambient)

% = reservoir temperature (ambient)

t = time

"N^p = compressor swept volume (total)

'\6 = dead volume between compressor and pulse tube

\ = volume of reservoir

M = volume of pulse tube

U = total internal energy

u = specific internal energy

W = rate of work input

X = position along the tube

Zf = flow impedance factor

oc = phase angle between liic and liicp

= phase angle between liih and riic

(j) = phase angle between in and mc

ip = phase angle between Td and m

1/ = linear frequency

u) = angular frequency

7 = ratio of heat capacities (Cp/Cy)

T = period of one cycle

Pa, = average gas density

7) = thermodynamic efficiency relative to Carnot

//c = viscosity at cold end

IV
CONTENTS

Page

NOMENCLATURE iii ABSTRACT 1

CHAPTER 1 INTRODUCTION 2

1.1 Purpose 2

1.2 Applications 3

1.3 Reliability of the Pulse Tube Refrigerator 4

1.4 Description of the Pulse Tube Refrigerator 4

1.5 AC Circuit Analogy 7

1.6 History of Pulse Tube Refrigeration Research 9

1.6a Thermal Acoustic Oscillations 9

1.6b The Basic Pulse Tube Refrigerator 9

1.6c The Resonant Pulse Tube Refrigerator 10

1.6d The Orifice Pulse Tube Refrigerator 11

CHAPTER 2

ANALYTICAL MODEL 13

2.1 Enthalpy Flow Analysis . 13

2.2 Phasor Analysis 18

2.3 Phase Angles 19

2.4 Mass Flow Rate Equation 20

2.5 Dynamic Temperature Equation 25

2.6 Enthalpy Flow Equations (Models 1 and 2) 27

2.7 Dead Volume in Pulse Tube Systems (Model 3) 32

CHAPTER 3

EXPERIMENTAL INVESTIGATION 35

3.1 Test Apparatus . , 35

3.2 Measurements and Calculations 36

3.3 Effect of Pressure on Performance 41

3.4 Effect of Working Fluid on Performance 42
3.5 Effect of Frequency and Mass Flow Ratio on Performance 46

3.6 Measurements of Gas Temperature Profile

in the Pulse Tube 49

CHAPTER 4

DISCUSSION OF RESULTS 55

4.1 Assumptions in the Model 55

4.2 Describing the Refrigeration Process 58

4.3 Optimizing the Orifice Pulse Tube Refrigerator 71

CHAPTER 5

CONCLUSIONS 73

ACKNOWLEDGMENTS 74

NOMENCLATURE 75

REFERENCES 77

APPENDIX A: DERIVATIONS 78

APPENDIX B: EXPERIMENTAL DATA 83

VI
Analytical Model for the Refrigeration Power of the

Orifice Pulse Tube Refrigerator

Peter J. Storch, Ray Radebaugh, and James E. Zimmerman

Chemical Engineering Science Division

National Institute of Standards and Technology

Boulder, Colorado 80303

The pulse tube refrigerator is a cryocooler with potential for high reliability. This

regenerative refrigerator operates with low average pressures, small compression ratios, and

only one moving part. In research conducted at the National Institute of Standards and

Technology, a pulse tube refrigerator with an orifice located at the closed, hot end of the

system has been used to reach 60 K in one stage. A gross refrigeration power of 12 W has

been measured at 80 K with a pulse tube volume of 30 cm^. This paper presents the

development of an analytical model to predict the refrigeration power produced in an

orifice pulse tube. Three versions of the model are presented: the first calculates

quantities which are independent of dead volume; the second neglects dead volume; and the

third describes a system with dead volume. Experimental results confirm predictions by

the model that refrigeration power is proportional to the average pressure, the pulse

frequency, the mass flow ratio, and the square of the dynamic pressure ratio. Results are presented that are useful for approximate design of orifice pulse tube systems. The model

assumes adiabatic conditions and an energy balance is performed in terms of net enthalpy

flow. Phasor analysis was found to be very useful for analyzing the oscillations of pressure,

temperature, and mass flow rate.

Key words: analytical model; cryocooler; cryogenics; pulse tube refrigerator; refrigeration

power; refrigerator.
CHAPTER 1

INTRODUCTION

The pulse tube refrigerator (PTR) is a new member of a family of devices known as cryocoolers. Cryocoolers are refrigerators which generate refrigeration at temperatures

below 120 K (—244*^) [1]. Pulse tube refrigeration is a cycle of compressions and

expansions which pump heat from one end of a tube to the other. Recently, there has been

a surge of interest in this method of refrigeration for use in applications which require

small, durable cryocoolers. The great advantage of the PTR is its simplicity, which

promises to provide high reliability. The lack of reliability in small cryocoolers is a well

known problem today.

1.1 Purpose

The general goal of this study is to gain a better understanding of the refrigeration

mechanism and the capability of the PTR. More specifically, this work is concerned with

developing an analytical model for the refrigeration power generated in an orifice pulse

tube refrigerator (OPTR). With the aid of several simplifying assumptions, we derive

equations which identify the important parameters in the system. Once the analytical

expressions are devised, verification of the model is accomplished by comparing results to

previous experimental studies as well as to experiments completed in this work.

A simple, analytical model of the cooling power is useful for three reasons. First,

such a model provides insight into the refrigeration mechanism. Often when a numerical

model is used to rigorously describe a system, it is difficult to see the physics involved in

the process. With an analytical expression, important parameters that affect performance

are more easily recognized. Second, basic trends or dependencies of performance can be

determined. Such dependencies will aid in optimization and indicate the range of

parameters which should be studied in more detail either numerically or experimentally.

Third, a simple analytical model will be useful for design. Although not as accurate as a numerical analysis, the model will be convenient for making approximate predictions of

system size and refrigeration capacity of the PTR.
1.2 Applications

Cryocoolers including the OPTR have extensive applications today and promise to

play an important role in future technology. Presently, cryocoolers can be found cooling

infrared detectors to temperatures below 80 K for sophisticated military aircraft, missiles,

and tanks. Parametric amplifiers used in ground—based systems for satellite

telecommunications and deep—space tracking typically require temperatures below 20 K

[2].- The most common commercial application of cryocoolers today is in cryopumping

which is important to the semiconductor industry [3]. Cryocoolers have also been tested in

space to cool gamma—ray and infrared telescopes [4]. For example, four Stirling

refrigerators, developed by a company in cooperation with Johns Hopkins University, flew

in space for approximately two years to cool two gamma—ray detectors.

The most important future application for cryocoolers such as the PTR is the

cooling of superconductors. It is increasingly apparent that superconducting devices will

greatly impact future technology and that cryogenics is the enabling technology. SQUIDs

(superconducting quantum interference device) typically require operating temperatures as low as 8 K. However, since 1986 there has been great progress in the development of high

temperature superconductors. New copper oxide materials show transition temperatures in excess of 100 K, and may have practical application at 40 or 50 K. Such temperatures are within the range of a two—stage pulse tube refrigerator. SQUIDs are used for making

geomagnetic and neuromagnetic measurements and have defense applications in submarine

and mine detection. Superconducting magnets at 20 K are used for magnetically

suspended, high speed trains, giant particle colliders, energy storage rings, and whole—body

magnetic resonance imaging (MRI). In addition, Josephson junctions may define a new

generation of digital circuits for high speed computers.

Nowhere is the need for reliabihty in small cryocoolers more evident than in space

applications. Research in cosmic—ray physics, plasma physics, and infrared sensor technology require cryogenic temperatures [5]. Most space missions today are planned to

last 3 to 10 years, so a cryocooler in space must operate reliably for a minimum of 3 years

to be useful and 5 to 10 years to be considered of real value. Therefore, successful

cryogenic designs for space applications require compact and reliable refrigerators with

extremely long lifetimes.
1.3 Reliability of the Pulse Tube Refrigerator

The PTR refrigerator has potential for both small size and high reliability. There

are two reasons for this potential. First, the pulse tube operates on a regenerative cycle as opposed to a recuperative cycle such as the Joule—Thomson (J—T) refrigerator. In a

regenerative refrigerator, the working gas oscillates between the piston at the hot end and

the cold end, where cooling is produced. As gas flows toward the cold end, heat is stored

in a unique type of heat exchanger known as a regenerator. Heat is retrieved by the gas as

it flows back to the hot end of the tube. Isothermal processes are used for heat addition

and rejection. Unlike refrigerators that use recuperative cycles, regenerative refrigerators

have neither a J—T valve to become plugged with impurities nor a low—temperature

expansion engine with high—pressure seals and valves to wear out. Yasukochi [6] has

shown that reciprocating expansion engines account for approximately 25% of failures in

recuperative cryocoolers. Regenerative cycles also have the advantages of a simple

compressor and a low operating pressure. J—T refrigerators require a valved compressor

with compression ratios of up to 200:1 and pressures as high as 20 MPa. The PTR,

however, uses only a basic piston without valves and typically operates with a pressure

ratio less than 2:1 and an average pressure less than 2.0 MPa. A simpler compressor with

lower pressures means less wear, less chance for a leak, and more dependable operation.

The second reason for the potential for high reliability in the PTR is its simplicity.

There is only one moving part—the compressor piston at room temperature. The lifetime

of mechanical refrigerators can be improved by eliminating moving parts. For instance,

the Stirling refrigerator, another regenerative refrigerator, has only two moving parts, and

is widely used today. The Stirling refrigerator shown in figure 1—1 operates with a piston

at ambient temperature and a displacer between the ambient and cold end temperatures.

However, the motion of the displacer generates two problems: the need for seals at low

temperatures, which tend to wear, and vibrations close to the refrigeration load, which can

decrease the lifetime of the instrument and disrupt the operation of a sensitive detector.

The PTR, with no moving parts at low temperature, has potential for longer lifetimes and

should be more compatible with sensitive instruments than the Stirling refrigerator.

1.4 Description of the Pulse Tube Refrigerator

Three versions of the pulse tube refrigerator are described in the literature today:

1) the basic pulse tube (BPTR), 2) the orifice pulse tube refrigerator (OPTR), and 3) the

resonant pulse tube refrigerator (RPTR). This designation was proposed by Radebaugh
Piston

Regenerator

Figure 1—1. Stirling refrigerator.

and coworkers [7] and will be used in this work. The subject of this study is the recently

introduced OPTR because of its promise of a refrigeration capacity higher than those of the

other two. The components and basic operation of the first two types of PTR are discussed

below, while the third is discussed later.

The three primary components of the BPTR, shown in figure 1—2(a), are a piston, a thermal regenerator, and the pulse tube itself. The regenerator is a periodic—flow heat

exchanger used to thermally isolate the piston at room temperature and the tube at the

cold temperature. The matrix of the regenerator is constructed of material with high heat

capacity to absorb heat from and reject heat to the gas as it flows back and forth. The

pulse tube is closed at the hot end where a heat exchanger transfers heat from the gas to

the surroundings. The tube is open at the cold end to the regenerator and the compressor

and has a heat exchanger to transfer heat from the refrigeration load to the gas. The

OPTR is shown in figure 1—2(b) with the additional components of an adjustable orifice

and the reservoir volume.
Regenerator

Piston

(a)

Heat

Exchanger

Pulse Tube —Qc

Heat

Exchanger

Reservoir

Volume

Piston

Regenerator

(b)

iii m iiii|iil Exchanger

Pulse Tube

—Qc

Heat

Exchanger

Figure 1—2. (a) Basic pulse tube refrigerator

(b) Orifice pulse tube refrigerator

The refrigeration cycle for both types of the PTR is best understood by examining

the temperature—position history of a single element of gas in the BPTR. Figure 1—3

shows the temperature of a gas element as a function of position in the tube. The first half

of the cycle begins when the piston moves gas into the tube. The gas is cooled to the

temperature of the cold end (point 1) as it flows through the regenerator. The element of

gas in the tube is compressed adiabatically and is heated as it travels toward the closed end

(point 1 to point 2). Constant—volume cooling then occurs during a relaxation period at high pressure (point 2 to point 3). During this period, heat is rejected in the hot—end

exchanger while the gas element in the tube is cooled by transferring heat with the tube

wall. The second half of the cycle occurs as the piston moves gas from the tube back

through the regenerator. An element in the tube is expanded adiabatically and cools as it moves toward the open end (point 3 to point 4). The cycle is completed with a relaxation

period of constant—volume heating at low pressure (point 4 to point 1). Gas in the
Wall

Temperature

POSITION

Figure 1—3. Temperature as a function of position for an element of gas in

the basic pulse tube refrigerator.

cold—end heat exchanger absorbs heat from the load while gas in the tube is heated by the

wall. The consequence of operating a system on this cycle is a net flow of enthalpy or pumping of heat from the cold end to the hot end. A constant temperature gradient in the

tube is established, and continuous refrigeration is produced.

The OPTR follows the same refrigeration cycle as described above except for the

method of heating and cooling during the relaxation periods. Instead of exchanging heat

with the tube wall, an element of gas in the OPTR experiences a change in temperature

due to flow through the orifice. As gas flows from the tube through the orifice and into the

reservoir, gas remaining in the tube expands adiabatically and cools. As gas flows back

into the tube from the reservoir, gas in the tube is adiabatically compressed and heats up.

1.5 AC Circuit Analogy

The oscillating behavior of flow in pulse tube systems is reminiscent of a common

AC electrical circuit. In fact, each mechanical component and action in the PTR has an

electrical counterpart. For instance, the flow of electrons or current in an electrical circuit

is analogous to the mass flow of the fluid in the pulse tube, while voltage is analogous to
pressure. A voltage drop occurs from resistance in a circuit just as a pressure drop occurs from anything that impedes flow in the PTR. Also, a capacitor, which stores charge, can

be used to represent a volume which stores mass in the PTR system.

The analogous circuit is shown in figure 1—4. The AC voltage source (element 1)

produces a sinusoidal voltage and current in the electrical system in much the same way

the compressor produces sinusoidal pressure and mass flow oscillations in the pulse tube.

However, for large pressure oscillations, there is significant deviation from sinusoidal

behavior and the compressor behaves more like a variable capacitor. Current from the

voltage source flows through the first resistor (element 2), which represents the solid

matrix in the regenerator, and a drop in voltage or pressure occurs. At this point, current

is diverted in the circuit to charge a capacitor (element 3), just as part of the mass flow

from the compressor is required to pressurize the regenerator void volume. The resulting

current then encounters the next two resistor—capacitor sets (elements 4, 5 and elements 6,

7), which represent dead volume and pulse tube volume with similar effect. Each volume

in the pulse tube can be associated with a resistance and a capacitance in the circuit. The

last resistor in the circuit (element 8) models the flow impedance of the orifice at the hot

end of the pulse tube. Again, a drop in voltage or pressure occurs here as charge or mass

flows into the large capacitor (element 9) representing the reservoir volume.

Drawing the analogy between the pulse tube and a familiar system such as a basic

AC circuit can help us better understand the function of components in a PTR. The

Figure 1-4. Analogous A C circuit for the orifice pulse tube refrigerator.
electrical model is also useful because methods for analyzing such systems are well

established. Specifically, phasor analysis and complex algebra, which are commonly used

to study AC circuits, can be applied effectively to PTR systems. Also associated with

sinusoidal voltage and current are phase angles; and it becomes very convenient to describe

oscillations in mass flow rate, pressure, and temperature with phase relationships.

1.6 History of Pulse Tube Refrigeration Research

1.6a Thermal Acoustic Oscillations

Heat pumping found in the PTR is closely related to another phenomenon known as thermal acoustic oscillations. Such oscillations occur spontaneously in tubes with a large

temperature gradient between a closed hot end and an open cold end, and they are commonly encountered today in the storage and transfer of cryogens. Successful theoretical

models for helium were presented by Rott in 1969 [8] and 1973 [9]. One of the earliest and

most well known observations of thermal acoustic oscillations was made by C. Sondhauss

in 1850. He noticed a sound produced from tubes used by glass blowers. When a gas flame

was applied to a glass bulb at the end of a tube open to the room, air in the tube oscillated

producing a sound characteristic of the tube dimensions. Lord Rayleigh in 1878 offered an

explanation for this spontaneous occurrence by pointing out that oscillations occur if heat

is added to the air at the point of highest pressure and removed at the point of lowest

pressure [10]. In other words, energy in the form of heat is introduced at the closed end

and energy in the form of acoustic waves and heat is removed at the open end. This

phenomenon is simply the reverse of the process in the PTR where energy is introduced at the open end in the form of pressure waves and removed at the closed end as heat to

provide useful refrigeration.

1.6b The Basic Pulse Tube Refrigerator

The use of thermal acoustic phenomena to generate refrigeration was first proposed

by Gifford and McMahon in 1959 [11]. The first pulse tube refrigerator apparatus, along

with a theory of the refrigeration process, was described in detail by Gifford and

Longsworth in 1964 [12]. This early PTR was similar to the device described above and

shown in figure 1—2. Gifford and Longsworth concluded that only those elements of gas

traveling between the hot— and cold-€nd heat exchangers were responsible for the

refrigeration effect. This idea was later refuted by the same authors in a second paper

9
written in 1965 [13]. Gifford and Longsworth found that the PTR still operated below the

critical pressure ratio where gas does not make the complete passage between the heat

exchangers. A third paper entitled "Surface Heat Pumping" was published a year later,

giving a much better qualitative description of the refrigeration effect [14]. The main

points of this description have been reviewed in the explanation of the BPTR noted earlier.

All experiments up to this point had been performed on PTR systems that used

valved compressors to generate the pressure wave. Longsworth used such a valved PTR in

1967 to reach a low temperature of 124 K with no load [15]. He measured a gross

refrigeration of 5 W at this temperature, which was totally absorbed by inefficiencies such

as regenerator ineffectiveness and conductive losses. That same year, Gifford and Kyanka

reached 165 K using a reversible PTR, that is, a PTR with a piston compressor instead of a valved system [16]. The idea of using a piston as a driver was first mentioned by Gifford

and Longsworth in 1965 as a method of overcoming the large inefficiency introduced by

irreversible expansion through the pressurization valves. Gifford and Kyanka showed that

the reversible system had a work requirement of only 10% of that of an ideal valved

system. However, the refrigeration capacity of the reversible PTR was shown to be much

less than the valved PTR at low pulse frequencies. This is due to the fact that a piston

generates pressure oscillations that are approximately sinusoidal. A valved system,

however, generates more of a "square wave" oscillation, which provides more time for heat

transfer between the gas and the wall. Gifford and Kyanka pointed out that performance

of the reversible PTR increases with frequency and should have been comparable to that of

the valved system at higher frequencies.

Although work on the BPTR was certainly interesting and gave valuable insight

into the mechanism of refrigeration, the lowest temperatures achieved were still too high to

be of practical use. Most applications of cryocoolers are below 80 K. Therefore, interest in

pulse tube refrigeration diminished and no great progress was made until recently when

some new concepts were introduced.

1.6c The Resonant Pulse Tube Refrigerator

A new idea in pulse tube refrigeration was presented by Merkli and Thomann in

1975 [17]. They showed in principle that a tube closed at one end and filled with gas

oscillating at a resonance frequency could be used as a heat pump. They measured cooling

in the section of the tube with maximum velocity (pressure node) and heating in the region

10
of the velocity nodes (pressure antinodes). However, the refrigeration produced in the tube

was minimal. Then, in 1983 and 1985, Wheatley and coworkers published two papers in

which they described a resonant tube having an internal structure of plates [18] [19]. This

structure or "stack" of plates becomes a second thermodynamic medium (the first is the

working fluid) in which a net enthalpy flow is generated. A basic schematic of RPTR is shown in figure 1—5. The system consists of an acoustic driver at the open end of the tube,

the resonance tube with the second thermodynamic medium, and a gas spring at the closed

end. A regenerator is not required with this configuration because the hot end of the plates

is toward the driver. In one case, a stack of cloth—epoxy plates was used in the tube, and a low temperature of 195 K was measured at a frequency of 516 Hz and an average pressure

of 1 MPa. The most recent work on resonant pulse tubes or thermoacoustic refrigerators

has been done by Tom Hofler in 1986. His apparatus uses a polymer film roll structure in

the resonator. With the hot end at 300 K, this device reached a cold end temperature of

200 K. An efficiency equal to 18% of Carnot efficiency was obtained.

1.6d The Orifice Pulse Tube Refrigerator

The second major innovation for the PTR came in 1984 when the Russian scientist

E.I. Mikulin inserted an orifice at the closed end of the tube [20—22]. He called this an

"expansion" pulse tube because the orifice allows part of the gas to expand into a large

reservoir. The effect of the orifice on the refrigeration mechanism was a large increase in

the refrigeration capacity. Originally, Mikulin placed the orifice between the hot tube and

the heat exchanger at the hot end and used a valved compressor with air as the working

fluid. With a pressure ratio (Pmax/Pmin) of 2, a frequency of 15 Hz, and no thermal load, a

Acoustic Driver

Closely-Spaced Plates

Gas Spring

Volume

Figure 1—5. Resonant pulse tube refrigerator.

11
temperature of 105 K was achieved and 10 W of refrigeration were measured at 120 K.

The pulse tube in this case had a volume of 35.3 cm^. However, the paper does not provide

details as to the amount of gas that was leaving and entering the pulse tube through the

orifice during one cycle.

In 1985, Radebaugh and coworkers at the NIST, formerly the National Bureau of

Standards (NBS), modified Mikulin's configuration slightly and relocated the orifice on top

of the hot end heat exchanger [8]. This configuration has been designated as the orifice

pulse tube refrigerator and is the same apparatus used in this work as shown in figure

1—2b. Radebaugh's version uses a piston compressor with helium as the working fluid. For

a 30 cm3 pulse tube, after optimizing the orifice setting a gross refrigeration capacity of 18

W was measured. This measurement was taken at the low temperature of 60 K with Pmax = 1.24 MPa, Pmin = 0.71 MPa, and a frequency of 9 Hz. The mass flow rate through the

orifice was calculated to be about 50% of the flow rate at the cold end. The point was

made that frequency and pressure ratio were not totally optimized for these measurements.

Also, with improved regenerators and multiple stages, Radebaugh indicated it may be

possible to reach a temperature of 10 K [8].

Experimental work on the OPTR at NIST has continued and some very useful

results were published in 1986 by Radebaugh and Herrmann [23] [24]. Many important

conclusions about the operation and performance of the OPTR were presented.

Measurements of efficiency and refrigeration capacity were made with four different pulse

tube sizes. Measurements were taken with helium gas at an average pressure of 1 MPa and

a pressure ratio (Pmax/Pmin) of 2.6. The best performance was obtained with a 30 cm^ tube

at a frequency of 6 Hz, generating 15 W of refrigeration power at 100 K. The intrinsic

thermodynamic efficiency measured for this case was 40% of Carnot efficiency at the

optimum orifice setting which occurred when the ratio of mass flow through the orifice to

mass flow in the tube (rhh/rhc) was 0.35. The thermodynamic efficiency was defined using

the P—V work in the tube (see eq (3—8)). The data supported the idea that efficiency

decreases with an increase in flow through the orifice. Radebaugh and Herrmann also

showed that, although the thermodynamic efficiency of the BPTR (orifice closed) was

generally twice as high as for the OPTR, the refrigeration power produced with the orifice was as much as 10 times greater than the BPTR. The fact that the OPTR has a lower

thermodynamic efficiency and a larger refrigeration power than the BPTR indicates the

presence of a different mechanism for pumping heat. Radebaugh and Herrmann also point

out that efficiency of the OPTR is comparable to that of the Stirling refrigerator and

12
greater than that of the Joule—Thomson refrigerator [24]. Finally, another important

conclusion from this work was that refrigeration capacity depends on volume and not on

length or diameter of the tube, although there is certainly a minimum limit on these

parameters.

CHAPTER 2

ANALYTICAL MODEL

Since the OPTR was first introduced in the late 1970's by Tarasov and Mikulin,

only one analytical expression describing the refrigeration power has been published. This

expression was first presented by Mikulin and coworkers in the Soviet Union in two papers

in 1982 and 1983 [20], [21], and then in the United States in 1984 [23]. Mikulin developed an adiabatic model for the tube assuming ideal heat exchange and the absence of friction.

Three equations were presented for the average expanded gas temperature, the refrigeration

power, and the thermodynamic efficiency [22]. Unfortunately, it is not clear how the model

demonstrates the effect of flow through the orifice on the refrigeration power.

Like Mikulin's model, the analysis presented in this paper treats the ideal case of an

adiabatic tube. Support for the idea that processes in the orifice pulse tube are actually

adiabatic is found in experiments by Radebaugh and Herrmann which show that heat

transfer between the fluid and the tube wall actually decreases the performance [24].

However, in this case, the idea of enthalpy flow is used to describe the system. This model

also attempts to show the role of the orifice in the refrigeration mechanism, which has

already been established experimentally as an important factor in performance [23].

2.1 Enthalpv Flow Analvsis

In the past, a typical method of analyzing regenerative type refrigerators has been

to assume isothermal compression and expansion so that the heat transferred in one cycle is equal to the P—V work done. The most obvious example of this method is the Schmidt

analysis used to model Stirling refrigerators [1, p. 134]. Adiabatic modeling has also been

applied to Stirling engines, for example, by Finklestein with very useful results [1, p. 142].

Unfortunately, such adiabatic models have not produced closed—form solutions, which is the goal of this work.

13
In 1985, Radebaugh took a new approach to regenerative systems when he employed

an enthalpy—flow analysis to compare three types of PTRs [8]. This is the same simple,

but powerful, method commonly used in the analysis of steady—flow, recuperative systems

such as the Joule-Thomson refrigerator. The study of enthalpy flow in a regenerative

system is more complicated due to the oscillating flow. Temperature, pressure, and mass

flow rate are functions of time and position in real systems. However, the advantages of an enthalpy—flow analysis are that it eliminates work terms in the energy balance and creates

simple boundary conditions. P—V work is included in the definition of enthalpy: H = U +

PV where U is internal energy. The system described by the model and to be analyzed for

enthalpy flow is that shown in figure 2—1. It is emphasized that this work does not

attempt to model the entire pulse tube system with the compressor and regenerator.

Rather, the system of interest here consists only of the tube, which is assumed to be

completely adiabatic, and the hot— and cold—end heat exchangers, which are assumed to be

ideal and isothermal.

Figure 2—1 presents the heat flows to be considered in the energy balance: <S[> is the net enthalpy flow, Qc is the rate of heat flow in at the cold end, or the refrigeration

effect, and Qh is the rate of heat flow out of the hot end. <flr> is more appropriately

defined later.

Regenerator

<Hr>

Pulse Tube

I

-<H> <H> l==tx3

Reservoir

Orifice

'H

Figure 2—1

The OPTR system described by the model. Net heat flows considered

in energy balance are displayed.

14
The energy balance for figure 2-1 begins with the first law of thermodynamics for a steady—flow, open system,

Q = W + rti(hout-hin), (2-1)

where Q is the refrigeration power, W is the rate of work produced by the fluid, rii is the

mass flow rate in the tube, and h is the specific enthalpy of the fluid at the inlet or outlet.

We define enthalpy flow as

ft = itih. (2-2)

For a pulse tube refrigerator with oscillating flow, the energy balance is

<Q> = <W> + A<fi>, (2-3)

where all energy flows are averaged over one period of oscillation.

First consider the cold end of the system, shown on the left half of figure 2—1.

Because the system boundaries do not change (constant tube volume) and the P—V work is accounted for in the enthalpy term, the work flow W is 0. The rate of heat flow into the

cold end then is equal to the change in the enthalpy flow,

Qc = A<fi> = <fi> - <flr>- (2-^)

The term <ftr> is the average enthalpy flow into the system from the regenerator due to

regenerator ineffectiveness, conduction in the regenerator matrix and tube wall, and other

loss terms. Therefore, Qc represents the net refrigeration power available, while the

enthalpy flow in the tube <ft> is the gross refrigeration power.

15
The energy balance is then likewise applied to the hot end of the system in figure

2—1. Because flow through the orifice is assumed to be isothermal, the net enthalpy flow over one cycle for the gas leaving and entering the tube is zero. Therefore, the rate at which heat is rejected at the hot end is simply equal to the enthalpy flow in the tube or the

gross refrigeration power,

Qh = <fi>. (2-5)

The energy balances in eqs (2^) and (2—5) reveal three interesting points about the

PTR. First, a unique property of this refrigerator is that the gross refrigeration and the

net refrigeration can be measured simultaneously. Second, if loss terms are negligible, Qc = Qh, and the available refrigeration power can be measured at either end of the tube. It

is important to remember that the statement Qc= Qh represents only an energy balance on

the tube and not the entire refrigeration system. An overall balance would include the heat

rejected at the compressor. Third, given the ideal, adiabatic tube, there is no addition of

heat at any point in this region and the enthalpy flow must be constant.

The energy balances also show that the gross refrigeration power in the orifice pulse

tube can be modeled by developing an equation for the net enthalpy flow <fi[>; this is now

the task ahead.

The derivation begins with the equation

dh =

5h pdT +dhW T

dP, (2-6)

which shows the change in enthalpy as a function of T and P. For an ideal gas with

specific heat capacity at constant pressure Cp, it is a standard result of thermodynamic

theory that

16
W = and dh Cp = constant. (2-7)

Using these results, integration of eq (2-6) gives

h = h(To) + Cp (T-To), (2-8)

where To is some reference temperature. To in this case is assigned to be the average

temperature Ta of the gas at any point in the tube. The instantaneous temperature is defined to be

T(x,t) = Ta(x) + Td(x,t). (2-9)

Td is called the dynamic temperature, which is the amplitude of the temperature

oscillation and is a function of time and position. Therefore, specific enthalpy in eq (2—8)

becomes

h = h(Ta) + CpTd. (2-10)

Then multiplying all terms by m and using the definition of enthalpy flow.

fi = fi[(Ta) + mCpTd. (2-11)

Integration over one cycle of period r gives the average or net enthalpy flow. Because the

first term in eq (2—11) integrates to zero,

17
T

<tL> = (Cp/r) J liiTd dt. (2-12)

Thus far, it has been established for the ideal system that the average enthalpy flow

depends on the oscillations of mass flow rate and temperature, as shown by eq (2—12).

Both of these oscillations are assumed to be sinusiodal in time, and both are functions of

position along the tube. However, in order to satisfy the law of conservation of energy, the

average enthalpy flow along the adiabatic tube must be constant. An important test for

thermodynamic consistency in the model will be the cancellation of the postion dependency

in the product of rh and Td.

2.2 Phasor Analvsis

Equation (2—12) is a general expression for the average enthalpy flow in the OPTR

for an ideal gas. It shows that <H> depends on the interaction between the mass flow rate

and the dynamic temperature oscillations. Flow through the orifice affects the forms of

these oscillations and their interaction. Therefore, it is difficult to visuahze and model

such oscillatory behavior without some simplifications.

The first step taken in describing the oscillations in the OPTR is to assume that the

pressure oscillation is sinusoidal in shape, which means the mass flow rate and temperature

oscillations also behave as sinusoids. This assumption is good for small ratios of dynamic

pressure to average pressure (Pd/Pa<<l) and greatly simplifies the mathematics involved

in the analysis. The second step taken to help visuahze the interaction of the oscillations

of pressure, temperature, and mass flow rate is to represent the sinusoids as phasors.

Phasor analysis is commonly used to study currents and voltages in AC electrical networks

[25]. A phasor is a vector operating in "phase space" instead of in physical space. It has

magnitude and direction corresponding to the amplitude and phase angle of a sinusoid.

The phasor operates in the frequency domain and rotates at an angular velocity

corresponding to the angular frequency of the sine oscillation in the time domain. The

right side of figure 2—2 shows phasor A with magnitude | A | and angle 9. As time

progresses, the phasor rotates counter—clockwise and traces out the amplitude of the sine

oscillation on the left side of the figure. Since all the oscillations in the pulse tube have the

same frequency, their phasors will have the same angular velocity and maintain their

relative positions in the frequency domain. These vectors can then be manipulated by the

18
f(t) COS

TIME DOMAIN

f(t)=|A|cos(cdt+0)

FREQUENCY DOMAIN

A=|A|Z^

Figure 2—2. Sine oscillation f(t) is produced in the time domain by rotating

ptiasor A in the frequency domain.

ordinary rules of vector geometry. A phase angle does not affect the form of the oscillation,

but only fixes the origin in time.

2.3 Phase Angles

Phase angles are used to describe the phase relationship between the various

oscillations in the pulse tube. According to eq (2—12), the average enthalpy flow is determined by integrating the product of rh and Td over one cycle. If these quantities are written in terms of sinusoids, the integral has the form,

I = Jsina;t sin(a;t—^)dt, (2-13)

where ip is the phase angle between the functions describing rh and Td- Integrating eq

(2—13) over one cycle and dividing by r results in the following term:

19
I = ^ COS V. (2-14)

Since <fi> is proportional to I, <ft> is a maximum when ^ = and lii and Td are in

phase. Similarly, <fi> = when iti and Td are 90° out of phase.

In a physical sense, the angle iJj represents the relative time difference between the occurrence of a flow rate maximum and the occurrence of a temperature maximum, ip is

affected then by the relaxation periods of cooling and heating in the pulse tube. A shift

in phase in the BPTR takes place when the gas exchanges heat with the tube wall. In the

OPTR, a phase shift occurs when gas in the tube is expanded or compressed due to flow

through the orifice. Since the phase angle is an important parameter for enthalpy flow, it

is expected that the method of phase shift will produce a difference in performance between

the two pulse tubes. (See Describing the Refrigeration Process in chapter 4.)

2.4 Mass Flow Rate Equation

The development of an expression for the mass flow rate involves an energy balance on the tube as shown in figure 2—3.

P,VpM^ K^mi

x = x = L

Figure 2—3. Adiabatic pulse tube system with mass flows riic at the cold end and

riih at the hot end.

20
The system we attempt to model consists of an adiabatic tube with constant volume

Vt, pressure P, and a total mass of gas Mt. The mass moving in and out at the ends of the

tube rile and riih, maintain constant temperatures Tc and Th. We assume no mixing of the

moving gas elements with bulk gas in the tube, which means the only effect of these mass

flows in the model is to change the pressure.

The energy balance for an open system with non—steady flow is

3Y = Q - W - Aft, (2-15)

where U is the total internal energy. For the tube with constant boundaries without heat

transfer between the gas and the tube wall, W = and Q = 0. The energy balance then

becomes:

d(/?Vu)

^

+ mhhh-mchc =0, (2-16)

where p is the gas density, V is the system volume, u is the specific internal energy, and hh

and he are the specific enthalpies of the gas at the hot— and cold-end temperatures

respectively. Assuming an ideal gas,

u = CvT = Cv^, (2-17)

hh = CpTh, (2-18a)

and he = CpTc. (2-1 8b)

21
After substituting eqs (2—17) and (2—18) into the energy balance, and using \ for the

general system volume, the energy balance becomes

g*^= mhTh-mcTc. (2-19)

Defining 7 as the ratio of heat capacities Cp/Cy and rearranging eq (2—19) gives the result,

Wr^3T = '^kT!-'^c- (2-20)

The left side of this equation represents the change in mass in the tube —-Tp due to mass

flow in and out at the tube ends. Because of the temperature gradient along the tube and

the assumption of no mixing between the working fluid entering and leaving the tube and

the bulk of the fluid in the adiabatic tube, the total change of mass is written in terms of

the cold end temperature. Integrating the equation

dMt _ Vt dP

dT"~ TET^ar

would not give the true total mass because it was assumed that mixing does not occur in

the tube. Th/Tc accounts for the effect of mass leaving or entering the tube at different

temperatures. The final result is ^= mh^-mc. (2-21)

In terms of phasors the mass balance becomes

22
dM

i=KlhijX-Klc (2-22)

In order to avoid confusion in the representations of the oscillations of mass flow

rate, pressure, and temperature, phasors are represented by underlined upper case letters,

amplitudes are shown in just upper case letters, and functions in the time domain are written in lower case letters. The phasor diagram depicting mass flow rates in the orifice

pulse tube is shown in figure 2-4.

The phasor Kic is taken as the reference phasor and assigned the value cos lA. All

other phasors, in terms of phase angle, are described in relation to this Nlc. For example,

is the phase angle by which mass flow through the orifice ^h lags the mass flow at the cold

end Kic-

-sm

-cos cos

sin

Figure 2-4. Phasor diagram of mass flow rate in the orifice pulse tube.

23
The angle 9 is also the phase shift between Pd and Mc because laminar flow through

the orifice is proportional to the dynamic pressure. The sinusoid for the dynamic pressure

is written simply as

Pd(t) = Pd cos {ut-9). (2-23)

Pressure drop in the pulse tube is neglected and pressure is therefore a function of time but

not of position.

The angle between Mh(Th/Tc) and dMt/dt must be 90° , so dMt/dt is proportional

to the pressure derivative dP/dt, as shown in eq (2—20). Figure 2—4 then gives an

important trigonometric relationship for phase angle 6:

cos^ = (Klh/Klc)(Th/Tc), (2-24a)

or sin^ = y 1 - (]VIh/Klc)2(Th/Tc)2 . (2-24b)

Another important feature of the phasor diagram in figure 2—4 is the total mass flow

rate, Kl- Mass flow rate in the tube is a function of temperature and pressure and therefore

depends on time and position in the tube. Total mass flow rate IVI varies from Is/lc at the

cold end (x=0 in figure 2—3) to l^h at the hot end (x=L). On the phasor diagram, IVI leads

iVlh by the angle and lags Jvic by the angle 6—(l). In a physical sense then, represents

position along the tube. Mass flow rate at any point in the tube is given below with the

following boundary conditions:

m(x,t) = M(x)cos(a;t-^+0), (2-25)

where at x=0, (j) = 9 and Kl = Klc, (2-26a)

24
and where at x=L, = and M = Klh. (2-26b)

2.5 Dynamic Temperature Equation

Along with the mass flow rate, the dynamic temperature will be used to establish

the enthalpy flow in the pulse tube. The development begins with the same energy balance on the pulse tube shown in eq (2—15). This equation is written for a point x along the tube

and gives the result

d(^ + d^^ Q ^2-27)

Using the continuity equation,

t + s:t = ». (2-28)

and again the definitions of internal energy and enthalpy for an ideal gas (eqs (2—17) and

(2—18)), we obtain the following differential equation:

5T_ RT ^P RTrti dT

(2-29)

If a small pressure ratio, (Pd/Pa << 1) is assumed, AP/P and AT/T are small and the

ratio T/P is equal to Ta/Pa- Equation (2-29) is then easily integrated over time to arrive

at the phasor representation of the dynamic temperature in terms of the dynamic pressure

(Ed) and the total mass (M) moved past point x in one-half cycle.

25
Td = RT

Pd RT_^^a

XTT^^r^- (2-30)

The phasor diagram for Td is shown in figure 2-5, where ^ is the phase angle between Td

and the total mass flow rate £. The first term of eq (2-30) is expressed in the time

domain by substituting eq (2-23) for Pd. An expression for M in the time domain is found

by integrating the total mass flow rate eq (2-25),

m(t) = (]^/tj)cos(a;t-^+(;/^7r/2) (2-31)

-sm

-cos cos

sm

Figure 2-5. Phasor diagram for dynamic temperature and components.

26
The phase of M lags that of M by 90° and has a magnitude of iCl/o;. The time domain

representation of the dynamic temperature is

Td(x,t) = ^^ p Pd cos(a-#)-^ gia U

dT,

~5^ cos{uA,-6+(f>-7r/2). (2-32)

2.6 Enthalpy Flow Equations (Models 1 and 2)

The average enthalpy flow is defined according to eq (2—12) as the integral of the

product of iti(x,t) and Td(x,t). Since the second term of Td in eq (2-32) is 90o out of

phase with respect to lii in eq (2—25), the product of these terms, will integrate to over one cycle. Therefore, the only contribution to <t{> from the dynamic temperature is

^ = ^^- (2-33)

Substituting eq (2—33) into eq (2—12) gives enthalpy flow in terms of dynamic pressure,

<fl> = ^/^rtiPd dt.

' ia

(2-34)

Applying the time—domain expression for lii and I^ results in an integral of the product of

two cosines,

<^> = ^^UPd L [cos(a;t-^+</))cos(a;t-^]dt. (2-35)

Performing the integration yields an expression for enthalpy flow in terms of the

time—average temperature profile in the tube %, (x); the average pressure la ; amplitudes

27
of the mass flow rate and pressure Kl(x) and Fd ; and the phase angle between Kl and S , 0(x):

_iRt; <fl> =

^ ^

KiPd COS0. (2-36)

The phasor diagram in figure 2—i provides a simple, geometric expression for 0,

COS0 = (Klh/IVl)(']h /Tk ). (2-37)

^{x) and %, (x) cancel with the replacement of cos^ in eq (2-36),

<ti> = i R^hTh (Pd /Pa ), (2-38)

and <ti> does not depend on position x. Thus enthalpy flow becomes independent of

position, and energy is conserved.

Equation (2—38) represents the net enthalpy flow or the gross refrigeration power in

the OPTR. By rearranging this equation, several convenient dimensionless forms result.

The first dimensionless equation is

<fi>

<mc>RX

= |(Pd /Pa )(Klh/]Vlc)('Ih /% ), (2-39)

where <rtic> is the mass flow rate at the cold end averaged over 1/2 cycle. Equation

(2—39) shows that enthalpy flow per unit mass flow is proportional to the pressure ratio

I^ /Pa , the mass flow ratio Klh/lVIc, and the temperature ratio % /% . In terms of the

phase angle 9, this equation is simply

28
.^^^ = f

(Pd/Pa )C0S^. (2^0)

<mc>RX

The set of eqs (2-39) and (2^0) is called Model 1; this set is independent of dead volume

in the system. These equations are useful for comparing performance per unit mass of the

pulse tube to that of other cryocoolers.

For design purposes, an equation in terms of the swept volume of the compressor Vcp is

desired. If dead volume in the system is neglected and if the compressor is assumed to be

isothermal at ambient temperature % , the mass flow rate at the cold end is

Uc = Kicp = ^ McpCJ =

I

^^iv, (2-41)

and so

<mc> =

I ^c = 2

^^

(2-42)

since w = 2x1/. Mcp represents the mass of fluid in the compressor and the maximum flow

rate IVIcp occurs when one-half of Mcp has left the compressor. Substituting for <mc> in

eq (2—39) results in the dimensionless form,

p^V^ = |(Pd/Pa)(Nlh/Klc). (2-43)

Equation (2-43) is useful for determining the cooling power available from a given

compressor size. The refrigeration power for this case is proportional to the average

pressure, the compressor volume, the frequency, the dynamic pressure ratio, and the mass

flow ratio.

Another dimensionless enthalpy—flow equation can be developed for a pulse tube

29
with constant volume Vt- To do so, an expression relating the pressure ratio in the tube to

the volume of the tube must be derived. The derivation begins with eq (2-20) from which

the differential change in pressure due to mass flow at the tube ends can be derived. If the

relationships in the time domain for the oscillating mass flow rates dihc and driih are substituted in eq (2—20), integration over time t results in the following equation for

pressure:

p(t) = ^[TcKlcSinort - ThKlhSin(cjt-^)]. (2-44)

The integration is performed assuming Tc and Th to be constant, which means the change

in temperature for gas entering and leaving at the ends is small. The amplitude of the

pressure oscillation is the dynamic pressure I?i , and using phasor algebra, the following

equation is obtained:

Pd = ^

[TlUl + nUl - 2TcThKlcKlh cos^]i (2-45)

After we factor out ^c and Tc, and we use eq (2—24) for cos^, the dynamic pressure

becomes

Pd =^^lVIcSin^. (2-46)

Substituting for Kic using eq (2-41) gives the pressure ratio in terms of the compressor

volume and the pulse tube volume.

Pd /Pa = i (Vcp /\t ){%/%) sm9. (2^7)

30
The above equation is appropriate for a reversible, adiabatic process with zero dead volume

between the compressor and the tube.

After we use eqs (2^3) and (2-47), two dimensionless equations for enthalpy flow

for a given pulse tube volume can be written in terms of the dynamic pressure ratio and the

volume ratio:

and

^^= ^

(Pd /Pa )'(lClh/^c)(Th /% )[l-(]^h/^c)'(Th /% )Y^\ (2^8)

^^=I

(Vcp l\ )'(^h/]VIc) (1^ /Th )[1 - (^h/Klc)'Clh /X )Y'/'- (2^9)

The equations can be simplified by writing them in terms of ^, ^^= ^(Pd/Pafcos^, (2-50)

and <fl> ^^= 7 f

(Vcp i\ )\\ /Th )2sin^ cos^. (2-51)

Equation (2—50) shows that the dimensionless enthalpy flow increases with the square of

the dynamic pressure ratio. If this ratio is constant, the enthalpy flow decreases as the

heat capacity ratio 7 of the working fluid increases. According to eq (2—51), enthalpy flow

increases with the square of the volume ratio and, if this ratio is constant, increases as 7

increases. The set of eqs (2-43), (2-50), and (2-51) is called Model 2.

31
2.7 Dead Volume in Pulse Tube Systems (Model 3)

Dead volume in pulse tube systems is very important for design. Up to this point,

the model of refrigeration power has ignored extra volume between the compressor and the

pulse tube. In real systems, this volume exists in connective tubing and in the regenerator

and will significantly affect performance. In the ideal system with no dead volume, liic =

rhcp. In the real system, mass must be used to pressurize the extra volume, which means

mass flow at the cold end will be less than the flow rate at the compressor. The smaller

mass flow results in a lower pressure ratio and a loss of cooling power.

A new equation for mass flow at the cold end will now be derived for a mass balance on the dead volume between the compressor and the pulse tube. The system is shown in

figure 2-^, and the mass balance is

dMo/dt = Kicp - Mc- (2-52)

The phasors in eq (2—52) are displayed on the phasor diagram for the mass flow rate in

figure 2-4. Md is the mass of fluid in the dead volume which is assumed to be isothermal

at temperature % . Therefore,

^T^^-ivxcp-ivxc 1t = ^cp-|^c. (2-53)

P. Vp , Md

Figure 2—6. Isothermal dead volume system between compressor and pulse tube.

32
In terms of phasors, dead volume creates a phase shift, so Kic lags Kicp by the phase angle

a. The mass balance can be written in the time domain as

\& d

j^ -^ = KlcpCOs(a;t + a) - ]^cC0sa;t. (2-54)

Integrating eq (2—54) results in an equation for pressure,

_RTh p(t) = ^^

[]^cpsin(a;t -\- a) - Uc^muA]. (2-55)

Again, phasor algebra provides an expression for pressure amplitude:

Pd = ^

[^?P + ^l - 2KlcpKlcC0sa]i (2-56)

We make use of figure 2^, and applying a trigonometric identity and the law of sines,

cos Of is written in terms of ^,

cosa = (Kic +

^

sin^)/Klcp, (2-57)

where the amplitude of the change in mass in the dead volume is

dMD _ PdVD

dT"-in^'^- ^^ ^^^

Substituting eq (2-57) into (2-56) results in an equation containing the dynamic pressure

amplitude V^ and the mass flow rate amplitude Kic,

33
'PdVD .2 P^VD (f^ cj) Mp + ^l - 2Uc (^^ a;)sin^. (2-59)

Equation (2—59), along with an expression containing I^ derived from a mass balance on

the pulse tube volume shown in eq (2—53), are then solved simultaneously to produce

expressions for I^ /I^ and ^cPd /Pa =

I

(Vcp/VD ) sin^(V*)% (2-60)

and Kic = ^ Ucp (Vt/\6 ) (Th/Tc)/(V*)% (2-61)

where

V* = [(l/72)(Vt/\f) )'(Th/Tc)%{(2/7)(Vt/\6 )(Th/Tc)+l}sin2^]. (2-€2)

V* is considered to be an effective volume ratio for an OPTR System with dead volume,

mass flow through the orifice, and an adiabatic tube.

After we substitute eqs (2-60) and (2-^1) into eq (2-39) (using the definition

2

<mc> = - Mc), the dimensionless enthalpy flow for pulse tube systems with dead volume

is ^^=

^ f

(Pd /Pa )(M /\f) ) cos^/V*, (2-^3)

or

34
^^= \^ (Vcp/Vd)(\( /Vi) cosfein^/V*. (2-64)

Equations ( 2—63) and (2—64) are in the same form as that of eq (2—50) and (2—51)

and show enthalpy flow in terms of the pressure ratio F^ /I^ and the volume ratio Yq? /Vd

respectively. They represent Model 3.

CHAPTER 3

EXPERIMENTAL INVESTIGATION

3.1 Test Apparatus

The experimental pulse tube apparatus was designed to provide both convenient and

accurate measurements of important parameters for the calculation of refrigeration power

and mass flow rate. A schematic of the apparatus is shown in figure 3—1. The regenerator,

pulse tube, and two units called isothermalizers are contained within the cryostat, while

the orifice and reservoir are located outside at room temperature. A vacuum of 10"2 p^

(10"4 Torr) or lower is maintained in the cryostat, and the components within are surrounded by a copper radiation shield maintained at 77 K by an LN2 bath. The hot—end

heat exchanger of the pulse tube, where measurements of gross refrigeration are made, is

carefully insulated from the radiation shield with several layers of aluminized Mylar.

All experimental measurements were carried out using a stainless steel pulse tube of

length 130 mm and diameter of 19.1 mm (X = 31.9 cm^). The isothermal ends of the

pulse tube consist of two copper caps filled with 80—mesh copper screen to ensure good heat

transfer between the gas and the cooling water. The screen also helps to align the flow in

the tube and decreases turbulence which degrades the temperature gradient. The volume

of each of the hot—and cold-end heat exchangers is about 10% of the pulse tube volume.

Conductive losses along the stainless steel pulse tube were calculated to be less than 1% of

the refrigeration capacity.

The regenerator was constructed of a stainless steel tube 19 mm in diameter by 127

mm long packed with 1050 discs of 150-mesh, phosphor-bronze screen. The void volume

in this regenerator was 21.6 cm^ and dead volume in the system between the compressor

35
and the pulse tube was 50 cm3. The hot end of the regenerator had a water—cooled heat

exchanger made up of 80—mesh copper screens. This heat exchanger ensures that gas

entering the regenerator is at ambient temperature. The orifice consists of a 15—turn

needle valve and the volume of the reservoir located above the orifice is approximately

1000 cm3.

In previous work by Radebaugh, two isothermalizer units were inserted between the

regenerator and the pulse tube in order to isolate regenerator loss from intrinsic

refrigeration. The isothermalizers have a large thermal mass with high heat capacity to

prevent temperature oscillations. There is excellent heat transfer with the working fluid

which flows through two small, annular gaps 178 fim thick. In past experiments, the two

isothermalizers were maintained at the same temperature by cooling the first with N2

boil—off and heating isothermalizer 2 with an electrical heater. Heat removed from

isothermalizer 1 represents the regenerator loss, and heat supplied to isothermalizer 2

represents the net refrigeration power. In this work, isothermalizer 1 was not cooled since

only measurements of gross refrigeration at the hot end of the pulse tube were needed.

The reciprocating compressor was valveless and nonlubricated with a total swept

volume of 290 cm^. The variable—frequency drive used was capable of providing

frequencies up to 12 Hz.

3.2 Measurements and Calculations

Figure 3—1 indicates points in the apparatus where temperature and pressure were

measured. A variable reluctance pressure transducer was located at the hot end of the

pulse tube, near ambient temperature, to measure the absolute pressure oscillation. A

similar pressure transducer is used in the isothermahzer 2, where flow is laminar and

essentially isothermal, to measure pressure drop for calculation of the mass flow rate in the

pulse tube. Since the isothermalizer is at the cold end temperature, the temperature

dependence of the transducer sensitivity was determined in a previous experiment [23].

The sensitivity was found to vary by less than 13% between 300 K and 77 K so this

variation was ignored. However, the zero setting of the pressure transducer changed

significantly with temperature and was therefore reset for each new measurement. A third

pressure transducer of the piezoresistive type was used to measure the absolute pressure

oscillation in the reservoir volume. The uncertainty in each of the readings of the three

pressure transducers was ±1 mV (±1 kPa).

36
w

1 1 -

Regenerator

Isothermalizer-

#1

Isothermalizer-

#2

[m

N2 gas

Reservoir

Volume

m] [^ I I [m

T X ^-^.ii^^-^ThrottleJL T

% H^OX Valve--^ %

H,0

g)c mjocAP

^AT=0

m

LN,

.Cryostat

Top

-Pulse

Tube

FillVent

-Radiation

Shield

Figure 3—1. Schematic of test apparatus.

37
The three pressure oscillations, along with the temperature oscillation in the pulse

tube (see temperature experiment), were collected on a processing digital oscilloscope. The

oscilloscope is capable of reading voltage with 12—bit resolution and an accuracy of ±0.75

percent.

Absolute temperatures at the cold and hot ends of the pulse tube were measured

with diode thermometers accurate to ±1 K. The difference in temperature between the

water inlet and outlet in the hot end heat exchanger was measured with a type—

T

(copper/Constantan) thermocouple.

An equation for the mass flow rate <mc> at the cold end of the pulse tube was

developed to account for density and viscosity variations with temperature. Measured

values of the pressure drop AP and the pressure ratio 111 /R across the laminar flow

element in the isothermalizer are used as follows:

<rhc> = M.

. r P

h ^+"^ 1'J ^17^'

^P

(3-1)

where pa is the density of the gas at the average pressure and at T^ , ^c is the viscosity at % , and Zf is the flow impedance for the annular gaps in the isothermalizer. The factor Zf was determined experimentally from AP data at different flow rates. Equation (3—1) is

valid for laminar flow, and the derivation is shown in Appendix A.

Flow of both He and N2 through the isothermalizer remains laminar for mass flow

rates up to about 2.5 g/s. For helium, flow rates were never greater than 2 g/s; however,

the flow rates with nitrogen were much higher and turbulent flow was encountered.

Therefore, eq (3—1) could not be used for N2, and a trial—and—error procedure was

necessary to solve for the mass flow rate. This procedure is presented in Appendix A.

The mass flow rate through the orifice can be expressed simply as the amount of mass moved in one direction divided by one half the period. For an ideal gas, adiabatic

flow, and a small pressure oscillation in the reservoir, the average flow rate is proportional

to the change in pressure in the reservoir as follows:

38
' _ Amh _ 2APrVr /<. ^x

where Vr and Tr are the volume and temperature of the reservoir, APr is the difference

between the maximum and minimum pressures in the reservoir, R is the gas constant, and

T is the period. Because a significant amount of gas is moving through the orifice at a high

velocity, adiabatic flow is a good assumption. Also, the pressure variation in the reservoir

was very small (APr/Pr < 0.02), so eq (3-2) is valid for this pulse tube system. The

derivation is shown in Appendix A.

Gross refrigeration was found by measuring the flow rate m and the temperature

change AT of the water flowing through the cooling coil at the hot end of the tube

according to the equation

Q = liiCpAT, (3-3)

The measurements of gross refrigeration were accurate to ±7%.

In order to calculate the efficiency of the pulse tube, it is necessary to find the work

input rate W. Because the compressor used to drive the pulse tube was not designed

specifically for this work, it is oversized and inefficient. Therefore, the electrical power

input to the compressor motor is not a good measurement of the work input required by

the pulse tube. However, by calculating the ideal thermodynamic work input from

pressure measurements, losses associated with the regenerator and the compressor are bypassed and only the efficiency of the cooling process is measured. This is the best

method for comparing the efficiency of the OPTR to that of other refrigerators.

The input rate of ideal thermodynamic work is determined from the measurements

of the mass flow rate oscillation at the cold end mc(t) and the absolute pressure oscillation

in the pulse tube p(t). For a closed cycle the work input is

W = pdV, (3^)

39
where V is the volume of the gas in the compressor, and is derived from the gas density and

mc(t). The work input over one cycle is

W = 1„ pW

^

dt. (3-5)

The average rate of work input is found by dividing by the period,

<W>=ij p(t)^dt. (3-6)

The coefficient of performance

(COP)pt = <Q>/<W>, (3-7)

of the pulse tube can then be calculated using time—averaged values of Q and W, and

compared to the COP for a Carnot cycle to determine the thermodynamic efficiency

relative to Carnot efficiency:

V = (COP)pt/(COP)c = [<Q>/<W>][Tc/(Th-Tc)]. (3-8)

Here, rj is the thermodynamic efficiency only for the cooling process in the pulse tube and is

useful for comparison with other refrigeration cycles.

All calculations for mh(t), rnc(t), V(t), and W were performed on the processing

digital oscilloscope.

40
3.3 Effect of Pressure on Performance

The analytical model indicates that the average pressure and the ratio of dynamic

pressure to average pressure are two important parameters in describing enthalpy flow in

the orifice pulse tube. These pressure parameters have not previously been studied

experimentally. The objective of this experiment is to verify the dependency of

refrigeration power on Pa and Pd/Pa as predicted by the model. Equation (2-50) shows

that refrigeration power is a linear function of Pa and a quadratic function of Pd/Pa- This

equation is particularly convenient since dimensionless enthalpy flow is in terms of pulse

tube volume and not compressor volume. The working volume of the compressor in the

experimental apparatus was varied in order to obtain data at different pressure ratios.

Refrigeration capacity was measured on the same OPTR as described earlier. The

frequency and the cold end temperature were kept constant at 6 Hz and 200 K, and the

orifice valve was set at three turns unless otherwise noted. The first set of measurements

was made at average pressures of 0.5, 1.0, and 1.5 MPa, with liih/mc = 0.28 and Pd/Pa =

0.29. Care was taken to assure that the compression ratio in these measurements was

constant because, as will be shown, the refrigeration capacity changes rapidly with Pd/PaIn the second set of measurements, the average pressure was kept constant at 1.0

MPa. The pressure ratio Pd/Pa was varied from 0.9 to 0.39 with liih/mc = 0.26. Different

compression ratios were achieved by adjusting a bypass valve between the inlet to the

refrigerator and ballast volume.

The results of this set of experiments are compared to the analytical model

represented by eq (2—50). This dimensionless equation is expressed in terms of the pulse

tube volume, which is constant throughout this work. The model shows that refrigeration

power is proportional to the average pressure. The experimental results are shown in figure

3—2. Even though tests were conducted at only three different pressures, the data points

fall very close to the best—fit line, with a coefficient of correlation r^ = 0.99. The plot

clearly demonstrates the linear dependence of refrigeration power on the average pressure,

as predicted by the analytical model.

The results of experimentally varying the pressure ratio compare equally well with

the theoretical model. Equation (2—50) shows that the refrigeration power increases with

the square of Pd/Pa and figure 3—3 confirms this result. For this plot, r^ = 0.99 and the

41
best—fit curve shown by the solid line has an exponent of 2.10.

3.4 Effect of Working Fluid on Performance

The second experiment was designed to determine the effect of a different working

fluid on performance of the OPTR. Up to this point, helium has been used in all experiments. Nitrogen was chosen as the second working fluid, and refrigeration power was measured for comparison with the helium data. Recall that eq (2—51) of the analytical

model shows that dimensionless enthalpy flow increases with 7, the ratio of heat capacities,

for a constant ratio of compressor volume to pulse tube volume (Vcp/Vt). Equation (2—50)

shows that enthalpy flow increases with I/7 for a constant ratio of dynamic pressure to

average pressure (Pd/Pa)- In order to test eq (2—51), measurements with nitrogen were

taken at the same volume variation of the compressor (bypass valve completely closed) as previously used with helium. The pressure ratios for this case were Pd/Pa = 0.39 for

helium and Pd/Pa = 0.27 for nitrogen. Then, to test eq (2-50), the compressor volume

variation was decreased (bypass valve partially open) and refrigeration measurements were made with helium at the same pressure ratio as nitrogen. The average pressure was kept at

1.0 MPa with z/ = 6 Hz and Tc = 200 K as before.

40

30

LUo

Q.- 20

<

QCLUO

E 10

U-LU

T

Dependence of Refrigeration

Power on Average Pressure

Tc= 200 K

»S ^1 = 0-29

rho/rhc = 0.28

0.5 1.0 1.5 2.0

AVERAGE PRESSURE, MPa

Figure 3-2. Experimental measurements of refrigeration power as a

function of the average tube pressure.

42
40

CC

LUo

Q.

<

CC

LU

O

30

20 -

- lOh

00. L-LU

1 1 1 Dependence of Refrigeration

on Pressure Ratio

Po

1 wer

\ - 200 K / _ V = 6 Hz /

Pa = 1.0MPa / -

rfio/rfic = 0.26 /

-

/ o

_

- ^

o/

1 1 1

-

0.1 0.2 0.3 0.4 0.5

d a

Figure 3-3. Experimental measurement of refrigeration power as a function of the ratio of dynamic pressure to average pressure.

A typical sample of the data used for comparison of the two working fluids

(including the refrigeration power and important system parameters) is presented in table

3—1. Care was taken to keep the system parameters as nearly constant as possible for each

case. For instance, with nitrogen it was necessary to open the orifice valve to 10 turns in

order to match the mass flow ratio attained with helium at three turns of the valves. The

first set of helium data in table 3—1 (Q = 18.1 W) is for the case of constant Pd/Pa and the

second set of data (Q = 29.9 W) is for the case of constant Vcp/Vt43
Table 3-1. Experimental refrigeration power and system

parameters for N2 and He

Q

(W)

Tc

(K)

Pa

(kPa) (Hz)

Pd/Pa rhh/iiic 7

N2 14.8 200.0 1059 6.08 0.28 0.18 1.40

He 18.1 200.0 1052 6.10 0.29 0.24 1.67

He 29.9 200.0 1069 6.07 0.29 0.20 1.67

The experimental and analytical results are displayed in table 3—2 as ratios of

refrigeration power for convenient comparison. The first column in table 3—2 shows that

about 1.2 times more refrigeration capacity was measured for helium than for nitrogen with

constant Pd/Pa- Also, about twice the capacity was measured for helium with a constant

Vcp/Vt. The third column of Table 3-2 shows the ratios predicted by the model. For

constant volume ratio, eq (2—51) says Qup/Qpj = '^'Hp/'^N ~ 1-103; while for a constant

pressure ratio, eq (2—50) says Qup Qiu = 1^ /'^He ~ 0.838 when other system parameters

are held constant. Although values of Pa, v, Pd/Pa, and mh/mc for the different cases are nearly equal, the differences are large enough to affect the refrigeration ratios. Therefore

the experimental ratios were corrected for these differences and are shown in the second

column of Table 3—2. Calculations for values in Table 3—2 are given in Appendix A.

The corrected refrigeration ratio for constant Pd/Pa, 0.848, nearly matches the ratio

predicted by the model, 0.838. This indicates that eq (2—50) gives a valid description of

the dependence of refrigeration power on the working fluid. Since Qup/Qm = Tm /Tttp'

nitrogen produces a slightly larger refrigeration capacity than helium for the same pressure

ratio.

On the other hand, the corrected refrigeration ratio for a constant Vcp/Vt, 1.856, is much larger than the ratio 7rrp/7|sj = 1.193 as predicted by the model. However, we

suspected that, even though the same setting of compressor volume was used for both

nitrogen and helium, the effective volume ratio may not have been the same. A short

44
Table 3.2. Comparison of refrigeration ratios

((5He/0N,)EXP (0He/<!N2)EXPe i(iue'%,'>

Model

constant (Pd/Pa) 1-226 1.8481 0.838

constant (Vcp/Vt) 2.026 1.8561 1.103

iThese experimental values are corrected for slight differences between the

He and N2 runs in Pa, i', Pd/Pa? 3,nd rrih/mc

experiment was devised to test this idea.

The test consisted of measuring the pressure ratio in the pulse tube and at the

outlet of the compressor for both helium and nitrogen with the apparatus running at room

temperature. Because of the volume of the regenerator, two isothermalizers, and a

significant amount of volume in the connective tubing between the compressor and the

pulse tube, a significant pressure drop exists for flow between these two components.

Therefore we anticipate that the pressure ratio in the pulse tube would be less than that in

the compressor. The measurements shown in table 3—3 confirm this idea, and they also

show that the pressure drop was significantly greater for nitrogen than for helium.

The larger pressure drop associated with nitrogen is attributed to the fact that the flow for

nitrogen in the regenerator and isothermalizers was turbulent. Since nitrogen has a much

larger density and a slightly lower viscosity than helium, the mass flow rates for nitrogen were about five times greater than those for helium (see eq (3—1)). While flow for helium

was clearly laminar with a Reynolds number of about 800, flow for nitrogen became

turbulent with a Reynolds number of about 10 000. The large flow rate of nitrogen then

translates into a large pressure drop, as shown by the following equation:

AP = ^^l^^ ,. (3-9)

45
Table 3-3. Pressure ratios

(Pd/Pa) (Pd/Pa)

Compressor Pulse Tube

He 0.50 0.42

N2 0.54 0.35

The problem with a large pressure drop is that less gas volume is moved through the pulse

tube which in turn decreases the amount of refrigeration produced. The considerable

difference in volume of nitrogen moved compared to that of helium explains the high

experimental value for QHe/QN2 in table 3—2. Therefore it is not possible to directly

verify the dependence of refrigeration power on 7 for constant Vcp/Vt with the present test apparatus. However, since eq (2—50) and (2—51) are essentially different forms of the same

eq and (2—50) was verified, we think that (2—51) is also valid.

3.5 Effect of Frequencv and Mass Flow Ratio on Performance

Pulse frequency and mass flow ratio are two more parameters that needed to be

studied to further verify the analytical model. However, experimental work was not

necessary in this case since useful data were available from experiments previously

performed by Herrmann and Radebaugh in 1986 [23], [24]. These experiments were carried

out on the same apparatus as used in this study and under the same working conditions.

Three different pulse tubes were studied for which refrigeration power was measured at

different frequencies and different orifice settings for cold—end temperatures down to 80 K.

The data used for comparison with the model were taken for a pulse tube with different

dimensions but with the same volume as the tube tested here.

Refrigeration power as a function of frequency is plotted in figure 3—4. The data were carefully chosen to be sure that only the frequency was varying and that other

parameters. Pa, Pd/Pa, Tc, and riih/rnc, important to the refrigeration power were kept

constant. According to the model, the refrigeration power increases linearly with the

frequency. This is exactly the result seen in figure 3-4, in which three of the measured

46
60

Dependence of Refrigeration Power

on Frequency

2 4 6 8

FREQUENCY, Hz

12

Figure 3-4. Measurements of refrigeration power as a

function of pulse frequency.

data points fit the best—fit line with r^ = 0.99. The fourth point at 11.5 Hz deviated

greatly from the linear behavior; however this measurement was taken at a different Pd/Pa

and, once corrected, also conformed to the same linear dependence.

The fact that the refrigeration power exhibits a linear dependence on the frequency

indicates that the average enthalpy flow in the tube is constant over one cycle. For Tc =

150 K, Pa = 1 MPa, Pd/Pa = 0.39, and mh/riic = 0.32, an enthalpy flow of approximately

3 J/cycle was observed. The more cycles that can be completed in a time period the

greater the enthalpy flow rate. Theoretically then there is no optimum frequency. Of

course in practice this is not true since pressure drop and ineffectiveness in the regenerator

degrade performance at the higher frequencies. There is also a limit imposed by the flow

impedance of the orifice valve in which further increase in the frequency decreases flow

through the orifice.

. A second parameter that was studied using data from previous experiments is the

47
mass flow ratio mh/rric. Measurements were taken with the orifice valve open to 1/2, 1, 2,

and 3 turns giving a range of flow ratios from 0.2 to 0.75 at t' = 6 Hz and Pd/Pa = 0.4. The model says that Q has a linear dependence on mh/riic- The plot of experimental data

in figure 3-5 for Tc = 188 K, i/ = 6 Hz, Pa = 1 MPa, and Pd/Pa = 0.40 confirms this

behavior. The data provide an excellent linear fit with a coefficient of correlation of r2 =

1.00. It is puzzling that the plot seems to show that for riih/nic = (orifice closed), the

refrigeration power will be zero. Although small compared to refrigeration power with the

orifice open, cooling has certainly been observed with the orifice closed; this is due to small

but nonzero heat transfer between the gas and the tube wall.

60

60-

40-

CC

LJJ o

CL

z

o"^ 30-

<

111 O

CC

U.

UJ

cc

20-

10-

1 1 1 Dependence of Refrigeration

on Mass Flow Ratio

1 Power

Tc = 188 K

V = 6 Hz y

Pa = 1.0 MPa /-

Pd/Pa = 0.40 >/

'^ 1 1 1 1 0.2 0.4 0.6 0.8 1.0

M,/M^

Figure 3—5. Measurements of refrigeration power as a

function of the ratio of mass flow through the orifice

to mass flow due to the compressor.

48

\
3.6 Measurement of Gas Temperature Profile in the Pulse Tube

The temperature oscillation is important for understanding the cooling mechanism

in the OPTR. From the development of the model, the time—averaged gas temperature Ta

and the phase angle ip between the temperature oscillation and the mass flow rate

oscillation are important parameters in determining the enthalpy flow. Measurements of

the gas temperature profile in the tube would allow for the calculation of the actual

enthalpy flow as a function of position in the tube. Temperature measurements would also

provide useful information on the behavior of the dynamic temperature Td and phase shift

ip along the tube.

A probe was devised to sample gas temperatures at different points along the tube.

The probe consists of a type E, Chromel/Constantan (nickel—chromium/copper—nickel)

thermocouple attached to one end of a 3 mm diameter stainless steel tube. Since an

accurate measurement of the amplitude of the dynamic temperature is desired, a foil type

thermocouple, 0.0127 mm thick, was chosen to provide a fast response. The response time

was calculated to be 8 ms, which is an order of magnitude less than the heat transfer period

available during operation at 6 Hz. The thermocouple tip extends 1 cm from the end of the

tube, which was sealed with epoxy. The temperature probe was inserted into the pulse

tube through an 0—ring seal at the top of the cryostat as shown in figure 3-6. The gas

temperature could be sampled at any position along the tube during operation and

measurements were generally made at 1 cm intervals.

Four temperature profiles were measured with the orifice opened to three turns for

cold—end temperatures of 275, 200, 150, and 110 K. One profile was measured with the

orifice closed at 200 K. Each measurement consisted of approximately six cycles of the

temperature oscillation which were collected and stored on the digital oscilloscope. Care

was taken to make sure the system was at equilibrium when the data were collected and

also that Tc was the same for each measurement along the tube. The temperature profiles

are for helium at an average pressure of 1.0 MPa and a frequency of 6 Hz.

49
^Thermocouple

Y/ Leads

O-Ring Seal

Cryostat Top

Copper Screen

:m^w^

Pulse Tube

Cold End

I?

:— To Reservoir Volume

^^^^^

f^Hot End

3 mm S.S. Tube

^^;;>;>^^?^

Cooling Coil

Foil Thermocouple

Copper Screen

i To

Regenerator

Figure 3-6. Temperature probe with foil Chromel/

Constantan thermocouple.

50
400

UJ

CC

H

<

CC

UJ

LU

300

200

100

T

Temperature Profiles

in Pulse Tube

0.2 0.4 0.6 0.8 1.0

RELATIVE POSITION

Figure 3-7. Average gas temperature as a function of position

in the tube for different cold end temperatures with the

orifice open.

The first set of results obtained from these measurements is the time—averaged

temperature profiles. Figure 3—7 shows average temperature vs relative position in the

tube for the four cold—end temperatures with the orifice open. The circles represent

measurements taken with the thermocouple probe, while the squares represent

temperatures measured with the diode thermometers on the isothermal heat exchangers at the ends of the tube. The average gas temperature changes linearly over the middle 60% of

the tube. We anticipate that the true gas temperature at the hot end would be slightly

above the hot-end temperature measured by the diode due to the finite AT needed for heat

transfer. For the same reason, the gas at the measured cold end should be slightly cooler

than the cold—end temperature. However, the slopes of these profiles are great enough to

put the points just before the two ends well above the hot end temperature and well below

the cold end temperature. In contrast, figure 3-8 shows that the temperature profile with

the orifice closed is much flatter than that with the orifice open. For Tc = 200 K, the slope

of the profile with the orifice open is twice as large as the slope with the orifice closed. In

51
400

LU2 300

ID<

LU

CL

LU 200

100

Orifice Open

Orifice

Closed

Temperature Profile

in Pulse Tube

Tjj = 200 K

0.2 0.4 0.6 0.8 1.0

RELATIVE POSITION

Figure 3-8. Comparison of average temperature profiles in the tube with

the orifice open and the orifice closed.

order to show that the profiles all have basically the same shape, a normalized temperature

was plotted as a function of position for each case in figure 3—9.

The theoretical and experimental temperature oscillations at different points in the

pulse tube are presented in figure 3—10. The theoretical oscillations were calculated using

eq (2—32) and show the sinusoidal behavior assumed in the model. The temperature

amplitude predicted by the model basically increases linearly from the cold end to the hot

end. However, the experimental temperature exhibits a large amplitude at the center

of the tube with much smaller amplitudes at the ends. This behavior is attributed to the

isothermal heat exchangers which fix the end temperatures. The oscillation at the cold end

in figure 3—10a has very broad peaks with amplitudes less than half of those of the

theoretical oscillations. The oscillation at the hot end in figure 3—10a has maxima with

sharp peaks with about the same magnitude as the sine oscillation, while the minima are much flatter. The temperature oscillation for the center in figure 3—10b shows considerable

deviation from sinusoidal behavior with sharp maxima and broad minima.

52
I

I

1 1 1 1 Normalized Temperature Profiles

1.0 - Cold End Temp,

o 110 K

o -

o o 200 K Open

t-\ ? °J

0.8 - Q 200 K Closed

^ 275 K

1

A

O

D

D

1

0.6

d

§

a

0.4 go

O

o

Q O 0.2

n a8 1 1

1 1 0.2 0.4 0.6 0.8 1.0

RELATIVE POSITION

Figure 3-9. Normalized temperature as a function of position in the

pulse tube where Tl and Th are the low and high values

respectively of the data set.

The temperature oscillation at the center of the tube is also shown in figure 3—10b

for operation with the orifice closed. Theory predicts that the amplitude for this case will

be slightly smaller than for the case with the orifice open since the average temperature

and the temperature gradient are smaller. The experimental amplitude for operation with

the orifice closed is about four times smaller than that with the orifice open.

53
400 p Hot End Tc= 200 K

120 240 360 480 600 720 840

TIME, ms

(a)

350r- Center T^= 200 K

/Experimental

-^Theory

Theory

120 240 360 480 600 720 840

TIME, ms

(b)

Figure 3-10. Comparison of experimental temperature oscillation with

theoretical sine oscillation at (a) cold and hot ends with

orifice open, and (b) center of tube with orifice open and

orifice closed.

54
CHAPTER 4

DISCUSSION OF RESULTS

4.1 Assumptions in the Model

In this analytical development, several basic assumptions are used to simplify the

analysis. Some of these assumptions are very good for typical pulse tube systems while

others place some serious limitations on the resulting enthalpy flow equations. A

comparison of theoretical and experimental values of refrigeration power reveals that

values calculated from Model 1 (eq (2-40)) are between 3 and 6 times greater than gross

refrigeration measurements. Figure 4—1 shows a plot of the ratio of theoretical to

experimental refrigeration power as a function of cold-end temperature. Similar plots are shown for Model 2 (2-50) and Model 3 (2-63), in figures 4-2 and 4-3. The ratio of

theoretical to experimental values for the model which neglects dead volume is between 8

and 12 while the ratio for the model which accounts for the effect of dead volume is between 3 and 6. After the various assumptions in the model are examined, an explanation

is offered for the difference between the experimental and theoretical values of refrigeration

power.

12

<

CC

Q.

X

LU

*>* >-

oco

LU

I

8-

4-

1 1 1 1 1

Model 1

1

o

o

o o

1

1 1 1 1 1-

1 80 100 120 140 160

T. (K)

180 200 220

Figure 4—1. Ratio of theoretical refrigeration power to experimental

values as a function of cold-€nd temperature for

<fl>/<Klc>RTc.

55
Best—fit lines are plotted in figures 4—1, 2, and 3 to show the general trend of

increasing divergence of the theoretical and experimental values with decreasing

temperature. The uncertainty in the refrigeration power ratios is ±7%.

The first assumption made in the model is that the working fluid is an ideal gas.

The typical gas used in the pulse tube is helium, which behaves ideally at temperatures

above 50 K. Ideal behavior is expected for the system in this work since temperatures

below 100 K are not encountered. The next assumption made is that the pressure

oscillation is a function of time only. This simplification is valid in this case since the pulse

frequencies used were not high enough to cause pressure fluctuations along the tube. The

model also describes a system at equilibrium, which means that both the cold-end

temperature and the average temperature gradient in the tube are assumed to be constant.

These conditions would not exist during system cool—down.

The assumption of an adiabatic tube makes the model specific to the orifice pulse

tube. The absence of heat transfer between the gas and the tube wall means that cooling occurs only due to expansion in the tube from flow through the orifice, and that entropy is constant in the tube. This is the ideal case for the OPTR since heat transfer can actually

be detrimental to performance [23]. With operation at relatively high frequencies and a g

<

CC

Q.

X

LU

->. >-

CC

o

Lil

100 120 140 160 180 200 220

T. (K)

Figure 4—2. Ratio of theroretical refrigeration power to

experimental values for the model neglecting dead volume.

56
<

cc

Q.X

UJ

cco

LU

220

Figure 4-3. Ratio of theoretical refrigeration power to experimental

values for the model accounting for dead volume.

large tube diameter, the adbiatic assumption is reasonable since the thermal penetration

depth is proportional to l/V"^- For a system with helium at 300 K and operating at 6 Hz,

the thermal penetration depth is about 1 mm. The thermal penetration depth assuming a

flat plate is 7 = 1 /y/TrpCpV. This value is small compared to the 10 mm radius of the tube

used in this work. Therefore, the assumption of an adiabatic tube is appropriate for the

system.

The most important simplification in the model is the assumption of sinusoidal

behavior for the pressure oscillation. This assumption means the temperature and mass

flow rate are also sinusoids. The mathematics is then greatly simplified and phasor

analysis can be used to describe the oscillations. In the real system, the volume variation

in the compressor is the true sinusoid and not the pressure which varies as 1/V. However,

for small compression ratios (Pd/Pa<<l), the pressure approaches a sinusoidal behavior.

In the test apparatus, compression ratios of up to Pd/Pa = 0-44 were observed, which is probably the upper limit for small Pd/Pa- Therefore slight deviation from a sine oscillation

was expected and was observed for the experimental pressure. The pressure oscillation had

a sharper peak than a pure sine oscillation which contributes to the decrease in magnitude

of the experimental enthalpy flow. However, the behavior of the pressure oscillation affects

57
only the magnitude and not the dependencies of the enthalpy flow on the different

parameters.

Since there is significant dead volume of about 50 cm^ in the test apparatus, the

model which assumes no dead volume is not very realistic. A comparison of Figures 4—2

and 4—3 gives an indication of the effect of dead volume on the performance of the pulse

tube. The model which neglects dead volume predicts an enthalpy flow more than two

times greater than the model which accounts for dead volume. Data from the gross

refrigeration measurements as well as the calculated values from the models are shown in

Appendix B.

Probably the most limiting assumption in the model is that the tube is infinitely

long; therefore neglecting end effects. The sinusoidal behavior of the various oscillations as discussed above is actually valid only for the center of the tube. The temperature at the

ends of the tube are fixed by the isothermal heat exchangers, which flatten the oscillation,

and cause a significant deviation from sinusoidal behavior. The end effects are noticeable

in a comparison of the experimentally measured temperature oscillation with the sine

oscillation predicted by theory in figure 3—10a. The smaller experimental amplitudes at

the ends contribute to the difference between the theoretical and experimental refrigeration

power.

Since the enthalpy flow in the model is the gross refrigeration power (as shown in

the development of the energy balance eq (2—5)), assumptions neglecting losses such as those due to conduction, heat leaks, or regenerator ineffectiveness are not necessary.

4.2 Describing the Refrigeration Process

From the experimental measurements of the temperature oscillation, two very

useful plots which characterize the OPTR can be constructed . The first plot, presented in

figure 4-4, shows the average temperature profile in the tube along with experimental and

theoretical profiles of the dynamic temperature. The solid line follows the plot of the

experimental time—averaged temperatures in the tube for Tc = 200 K; the slope is very

steep for operation with the orifice open. The broken lines represent temperature

amplitudes for the system according to eq (2—32) of the model. The dashed lines show the

experimental amplitudes. As observed before, the actual amplitude above the average

temperature is very large in the center of the tube, much larger than predicted by theory.

58
Below the average temperature, the theoretical curve nearly matches the actual values.

400

LU

00

H

<

DC

LU

Q.

LU

300

T r

2009^cy

100

Tc= 200 K

Orifice Open

0.2 0.4 0.6 0.8 1.0

RELATIVE POSITION

Figure A—i. Comparison of dynamic temperature profiles to

the average temperature profile.

59
The large dynamic temperature and the steep average temperature profile can be

explained by looking at the gas in the tube center. Due to the motion of the piston, gas in

this region moves back and forth but does not leave the tube. As the gas moves, the

temperature gradient moves with it, alternately moving gas at the hot-end temperature

and then gas at the cold-end temperature to the middle of the tube. The greater gas

displacement with the orifice open than that with the orifice closed explains the steeper

temperature gradient observed in the OPTR as compared to the BPTR. It is understandable then why there exists an optimum orifice setting in the OPTR. The larger

the orifice, the greater the expansion of gas in the tube. However, if the orifice is too large,

gas will travel the entire length of the tube and the temperature gradient will deteriorate.

In effect, the quantity of gas in the center of the tube acts as a buffer between the two ends

preserving the temperature gradient. The buffer gas in the OPTR tube seems to act as a

"gas displacer" and to play a role very similar to the displacer in the Stirling refrigerator.

The second plot using the experimental temperature oscillation is shown in figure

4—5. Gas temperature is plotted against the mass of gas moved in the tube. The three

cases displayed here are for the cold end, the hot end, and the center. This figure is

referred to as a T—m diagram and is analogous to the P—V diagram commonly used to

study thermodynamic systems. Just as the area in a P—V diagram represents the work per

cycle in a system, the area in a T—m diagram represents the enthalpy transported per

cycle. Figure 4—5 gives a good description of enthalpy flow in the OPTR. The enthalpy

flow at the cold end is characterized by a large amount of mass with a very small dynamic

temperature. Enthalpy flow in the center of the tube is due to a smaller amount of gas but

with a very large temperature amplitude. The hot end has the least mass transported with

intermediate temperatures.

At this point, an attempt is made to describe the refrigeration process in the PTR

in both a physical and a thermodynamic sense. The BPTR and the OPTR operate on

similar thermodynamic cycles. The process consists of heating due to adiabatic

compression, a shift in phase between the mass flow rate oscillation and the temperature

oscillation, and then cooling due to adiabatic expansion. The experimental temperature

measurements, however, have shown that there is a significant difference in the average

and dynamic temperature profiles in these two refrigerators. Also, the difference leads to

greater refrigeration power in the OPTR. Figure 4—6 shows the temperature as a function

of position in the tube to help illustrate the difference in the refrigeration process. The

difference is in the method of phase shift. The BPTR relies on heat transfer to the tube

60
400

LU

CC

H

<

CC

LU

Q.

LU

290

0.10

MASS, g

Figure 4—5. T—m diagram characterizing enthalpy flow at the hot end,

center, and cold end of the pulse tube.

wall to cool the gas after the compression phase while the OPTR relies on the expansion of

gas in the tube due to flow through the orifice to cool the gas. The cooling and therefore

the phase shift in the BPT are limited to the temperature of the tube wall. Figure 4—6

shows how the gas in the OPT can be cooled below the wall temperature (2—3). During

flow through the orifice back into the tube (4-6), the gas is heated above the wall

temperature. This behavior is confirmed experimentally by the large dynamic

temperatures seen in figures 4—2 and 3—10b. The larger temperature amplitudes generated

by the orifice result in a larger enthalpy flow. In figure 4-6 the greater area enclosed by

the path on the left explains qualitatively why refrigeration power is greater in the OPT.

61
LU

CC

<

CC

LU

Q.

LU

Pulse Tube

Orifice

Tube Temperature

Heat

Transfer 2

Orifice Pulse Tube

Basic Pulse Tube

(orifice closed)

4 ^Expansion/Compression

due to flow through orifice

POSITION

Figure 4—6. Temperature as a function of position for an element

of gas in a pulse tube with the orifice open and with

the orifice closed.

Another method of describing the refrigeration process in the orifice pulse tube

involves a T—S diagram. It is impossible to represent pulse tube refrigeration on a T—

S

diagram in the traditional sense since each element of gas in the system follows a different

path. For example, the buffer gas in the center of the tube traces out a simple vertical line on the T—S diagram since it experiences only the constant entropy processes of adiabatic

compression and expansion. The gas which moves in and out of the ends of the pulse tube

is more interesting. Figure 4—7 shows a T—S diagram for an element of gas which enters

and leaves the cold end of the tube. Starting in the compressor at point 1 in the figure, the

gas is cooled to the cold end temperature as it is moved by the piston through the

regenerator and the cold end heat exchanger (1—2). The gas then experiences adiabatic

compression in the tube and follows path (2—3). At this point, the gas in the OPTR

follows path (3^) and is cooled by adiabatic expansion due to flow of other gas elements

62
ENTROPY S

Figure 4—7. T—S diagram representing the thermodynamic

cycle of an element of gas at the cold

end of the tube.

out the orifice and into the reservoir. The path corresponds to path (2—3) on the T—

diagram in Figure 4—6 where gas moves toward the hot end as it cools below the average

temperature. In the BPTR, however, cooling occurs at constant volume (3-4'), due to

heat transfer with the tube wall and is limited to the wall temperature. The piston then

moves gas back toward the cold end and the element of gas cools adiabatically from 4 to 5

in the OPTR and from 4'to 5' in the BPTR. The gas flows back through the cold-end

heat exchanger and heat is absorbed at constant pressure along the paths (5—6) and

(5'-6'). The gas completes the cycle back at point 1 in the compressor. The area under

the curves (5-6) and (5'—6') represent the refrigeration power generated in the OPTR and

BPTR, respectively. Just as shown in the T—x diagram in figure 4—6, the area is greater

for the OPTR than for the BPTR.

63
0.3

;^ 0.2

a

•I

\/ 0.1

1 1 1 1 /I

Vcp/VD=1.0 /

V\7>/

3^ y/^

^

^'^ 1 1 1

1 0.2 0.4 0.6 0.8 1.0

M,/M,

Figure 4—8. Gross refrigeration power as a function of

mass flow ratio for different temperature

ratios (Th = 350 K).

If the theory of operation described above in figures 4—6 and 4—7 is correct, we will

find that heat transfer in the OPTR is detrimental to performance. That is, heat transfer

would limit the dynamic temperature in the OPTR and decrease refrigeration capacity.

This is exactly the behavior observed by Radebaugh and Herrmann in their 1986 work.

They showed that the OPTR operated better at high frequencies where heat transfer is

insignificant [23]. Also, in this work, a significant loss of performance in the OPTR was

observed when the temperature probe was inserted into the tube. The farther the probe was inserted, the greater the loss of refrigeration power. (See data in Appendix B) The

loss of performance is attributed to the increase in heat transfer area.

64
0.20

0.16

(B

Q?

0.10

0.06

1 1 1 1

V{ / Vq U. 1 u

~ / 0. le^---^'''

/ / y^^^^

"

~

-

Th/Tc = 3

Vcp/VD-0.6

1 1

0.3

0.2-

Q?

0.1

oi

20 40 60 80

PHASE ANGLE, ^

(a)

\

1

1

^^^""VN^O.IO

1 \

—

Th/Tc =3

Vcp/Vd=0.83

-

20 40 60 80

PHASE ANGLE, Q

(b)

Figure 4-9. The effect of mass flow through the orifice on pressure ratio.

65
0.6

,co

Q?

Th/Tc=3

Vcp/Vo =1.5

20 40 60

PHASE ANGLE. 9

80

<a

QP

0.8

0.6

0.4

0.2

(c)

1 1 Vt/VD=OJO___—

- /^

0J^5_^- —

/ /^^-^S,^^——

7 / /O"""''^^^-'''-^

1 1

~~

'

-

Th/Tc = 3

Vcp/Vp = 2.0

1 1 20 40 60

PHASE ANGLE, B

(d)

Figure 4—9.

66

80
LU

Q

0.4 0.8 1.2 1.6

VOLUME RATIO, V^/V^

Figure 4—10a. Gross refrigeration power as a function of

pulse tube volume for six different compressor

sizes and four different cold end temperatures

(Th = 350 K).

67
a

>"

n QA

OB U.OO

Ql

•*^ I

0.32

\/

O

0.28

J

U0.24 >

Q.

1 < 0.20

H

Z 0.16

LU

CO

CO 0.12

LU

-1

z 0.08

o

CO

z 0.04

LU

2

Q

Figure 4—10b

68
c^

0.24

V,/Vd

Figure 4—10c

69
^ 0.20

a

>^

Q?

0.16 I

v

O

—I

LL

> 0.12-

Q.

<

I

IZ

LU 0.08-

CO

CO

LU

O 0.04-

CO

z

LU

Th/T,=4

e =30°

1

0.2 0.4 0.6 0.8 1.0

Figure 4-1 Od

70
4.3 Optimizing the Orifice Pulse Tube Refrigerator

For design the most useful expression resulting from this analytical study is eq

(2—64). This equation is written in terms of important design parameters including

compressor swept volume, pulse tube volume, dead volume, average pressure, pulse

frequency, and mass flow rate through the orifice or phase angle 9. We must realize,

though, that this equation describes only the pulse tube itself and does not account for the

effect of regenerator parameters on the optimum pulse tube configuration. An effort to

determine the optimum relationship between regenerator and pulse tube volumes was

conducted by Chowdhury and Radebaugh [26].

Figure 4—8 shows dimensionless enthalpy flow from eq (2—66) as a function of

mass—flow ratio for different cold-end temperatures. The gas temperature at the hot end

{\ ) was approximately 350 K for this system. For a given temperature ratio, refrigeration

power increases as flow through the orifice increases due to the expansion of gas in the

tube. However the effect of the expansion is offset by a decrease in the compression ratio

I^ /la , and a maximum in refrigeration power results. The point at which the maximum

occurs is the optimum mass—flow ratio and depends on the temperature ratio and the ratio

of pulse tube volume to dead volume. Of course the magnitude of the maximum is a

function of the compressor size or \^p /\6 . For figure 4—8,

Vcp /\i) =1.

The effect of mass flow through the orifice on the pressure ratio is displayed in figure 4—9. This series of plots for different compressor sizes (\{;p /X-f) ) and

pulse tube sizes {\ /\& ) show I?i /Pa as a function of phase angle 6. When the orifice is

closed, 6 = 90° and Pd /Fa is at a maximum. As the orifice is opened, IVIh/IVlc increases

while and I?j /Fa decrease. Plots 4—9a through 4—9d can be used to predict the

approximate pressure ratio obtained in a given pulse tube system for % /% =3. The

curves in these plots, constructed from eq (2—60), predict values of Ici /la about 5 times

greater than experimental measurements. The expression for Pd /Fa which neglects dead

volume, eq (2^7), predicts values about 10 times greater than experimental

measurements.

Another important series of plots is shown in figure 4-10. Dimensionless enthalpy

flow from eq (2-64) is plotted versus the volume ratio \ /\i) for six different values of

71
cp /\6 and four different cold end temperatures. These plots show a maximum in

refrigeration power and the optimum pulse tube volume can be determined for a given

compressor size, dead volume, and cold end temperature. For comparison, the optimum

\ /\f) for % = 80 K (Th /X =4) predicted by the model is 0.22 while the experimental

ratio was 0.20 [26]. This experimental ratio was calculated using the pulse tube system in

reference [26]: regenerator IV with \( = 7.35 cm^ and % = 81 K.

The plots shown in figures 4—10a through 4—lOd can be constructed for different

values of phase angle 9. The model, however, predicts that the maximum refrigeration

power occurs in the limit as the phase angle 9 and pulse tube volume \{ approach 0. In the

context of the model, this result is reasonable. The maximum expansion in the tube occurs when ^ = and the maximum pressure ratio occurs when \( = 0. However, lower limits on pulse tube volume and orifice opening size are defined for a real system by the need to

preserve the temperature gradient in the tube. As explained previously, performance

decreases when the so—called buffer gas leaves the tube, which occurs if the tube volume is too small or the phase angle is too small. The optimum phase angle for any pulse tube

system was determined to be 30° from the experimental data shown in the Appendix B.

One important point to keep in mind when applying the curves in figure 4—10 is that the value of \f) is dead volume at ambient temperature. Dead volume which exists at the cold end of a system must be corrected for temperature to obtain a realistic value of

72
CHAPTER 5

CONCLUSIONS

The analytical study presented here has contributed to our understanding of the

refrigeration mechanism and the capability of the pulse tube refrigerator. We have learned

that there are significant differences between the mechanism of the OPTR and the BPTR.

Heat transfer between the tube wall and the fluid is undesirable in the the OPTR since it degrades the dynamic temperature and therefore the gross refrigeration power. The BPTR,

however, relies on heat transfer to create the heat pumping effect. Clearly, flow through

the orifice generates greater refrigeration power and gives the OPTR more potential for

practical application that the BPTR.

The energy balance used in this analytical development reveals that both net and

gross cooling power can be measured in the PTR unlike other cryocoolers, . We have also

learned that the enthalpy flow analysis and the use of phasors are convenient and effective

methods of analyzing pulse tube systems. This approach can be generalized and should be

a useful tool for the study of other regenerative refrigerators.

An obvious test for validity of an enthalpy flow analysis is conservation of energy.

The enthalpy equations generated in this work satisfy this requirement since they predict

constant enthalpy flow along the adiabatic tube.

Experimental results also support the validity of the model. The dependence of

refrigeration power on average pressure, pressure ratio, frequency, and mass—flow ratio

predicted by the model matches experimental data. Furthermore experiments and the

model agree that helium produces the best performance when used for the working fluid for

the OPTR.

The refrigeration powers predicted by the model, however, are at best 3-6 times

greater than experimental measurements. The difference is attributed to the simplifying

assumptions made in the derivation of the model in order to obtain closed solutions. These

assumptions include laminar flow in the pulse tube as well as through the orifice, sinusoidal

variations in pressure, temperature, and mass flow rate, and an infinitely long tube. In

order to improve the analytical model and our understanding of the OPTR, further

experimental investigation is required. One recommendation for such work is a study to

73
characterize flow in the pulse tube to determine the effects of turbulence due to oscillating

flow and a tube of finite length.

The result of including dead volume in the enthalpy flow analysis is an increase in

the complexity of the equations and a factor of 2 decrease in the refrigeration power

predicted by the model. Dead volume is therefore an important parameter to consider in

the design of PTR systems.

The refrigeration power measurements taken in this work are consistent with

previous results. For example, at 200 K with a pulse tube volume of 30 cm^, a gross

cooling power of 34 W and an efficiency of 17% of Carnot efficiency were measured. These

values match results reported by Herrmann and Radebaugh in 1986 [23].

The plots generated from the model and presented in Chapter 4 are useful for

approximate design of an OPTR system. The system described here consists only of the

pulse tube itself. The regenerator is treated only as dead volume in the system. Before an

overall system can be effectively designed and optimized, the relationship between the

regenerator and pulse tube volume must be more clearly understood.

ACKNOWLEDGMENTS

We would like to express our thanks to Bill Horsly of Ball Aerospace Systems, and

David Smith and Kathy McDermott of NIST for their valuable technical and literary

input. Also, the expert graphics and word processing talents of Nick Sanchez and Marilyn

Yetzbacher are greatly appreciated.

74
¥

REFERENCES

[1] Walker, G., Crvocoolers Part 1: Fundamentals . Plenum Press, New York, 1983.

[2] Walker, G., Crvocoolers Part 2: Applications . Plenum Press, New York, 1983.

[3] Radebaugh, R., "Prospects for Small Cryocoolers," Proceedings of the 9th

International Cryogenics Engineering Conference. May 1982, p. 761.

[4] Naes, L.G. and Nast, T.C., "Long Life Orbital Operation of Stirling Cycle

Mechanical Refrigeration," Proceedings of the 8th International Cryogenics

Engineering Conference. IPC Science and Technical Press, Surrey, England, 1980,

pp. 204-212.

[5] Salomonoyich, Sidyakina, Khaikin, Bakun, Nikonoy, Maslakoy, Klimenko, and

Kurkin, "Space Helium Refrigerator," Cryogenics. August, IPC Business Press

Ltd., 1981, pp. 474-478.

[6] Yasukochi, K., Cryogenic Engineering, (in Japanese), Vol. 13, 1978, pp. 257—271.

[7] Radebaugh, R., Zimmerman, J., Smith, D.R., and Louie, B., "A Comparison of

Three Types of Pulse Tube Refrigerators: New Methods for Reaching 60 K."

Advances in Cryogenic Engineering. Vol. 31, Plenum Press, New York, 1986, p. 1979.

[8] Rott, N., "Damped and Thermally Driven Acoustic Oscillations in Wide and

Narrow Tubes," ZAMP, 20:230-243, 1969.

[9] Rott, N., "Thermally Driven Acoustic Oscillations, Part 2: Stability Limit for

Helium," ZAMP, 24:54-72, 1973.

[10] (Lord) Rayleigh, "The Explanation of Certain Acoustical Phenomena," Nature, London, Volume 18, 1887, p. 319.

75
[11] Gifford, W.E. and MacMahon, H.O., "A New Refrigeration Process," Proceedings

of the 10th International Congress of Refrigerations . Copenhagen, Denmark,

August 1959.

[12] Gifford, W.E. and Longsworth, R.C., "Pulse-Tube Refrigeration," Transactions of

the ASME, Vol. 63, 1964, p. 264.

[13] Gifford, W.E. and Longsworth, R.C., "Pulse-Tube Refrigeration Progress,"

Advances in Cryogenic Engineering. Vol. lOB, Plenum Press, New York, 1965, p. 69.

[14] Gifford, W.E. and Longsworth R.C., "Surface Heat Pumping," Advances in

Cryogenic Engineering . Vol. 11, Plenum Press, New York, 1966, p. 171.

[15] Longsworth, R.C, "An Experimental Investigation of PPulse Tube Refrigeration

Heat Pumping Rates, " Advances in Cryogenic Engineering. Vol. 12, Plenum Press,

New York, 1967, p 608

[16] Gifford, W.E. and Kyanka, G.H., "Reversible Pulse Tube Refrigeration,"

Advances in Cryogenic Engineering. Vol. 12, Plenum Press, New York, 1967, p. 619.

[17] Merkli, P. and Thomann, H., "Thermoacoustic Effects in a Resonance Tube," Fluid

Mech., Vol. 70, 1975, p. 161.

[18] Wheatly, J., Hofler, T., Swift, G.W., and Migliori, A., "An Intrinsically

Irreversible Thermoacoustic Heat Engine," Acoustic Society of America. Vol. 74,

1983, p. 153.

[19] Wheatley, J., Hofler, T., Swift, G.W., and Migliori, A., "Understanding Some

Simple Phenomena in Thermoacouystics With Applications to Acoustical Heat

Engineering," American Journal of Physics . Vol. 53, 1985, p. 147.

[20] Mikulin, E.I., Shkrebyonock, M.P., and Tarasov, A. A., "Investigatiion of Cryogenic

Pulse Tube," (in Russian) Higher Educational Institutions News. Machinery

Engineering. 1:71 (1983)

76
[21] Mikulin, E.I., Shkrebyonock, M.P., and Tarasov, A. A., "Expansion Cryogenic Pulse

Tube," (in Russian) Higher Educational Institutions News. Machinery Engineering. 9:76 (1982)

[22] Mikulin, E.I., Tarasov, A. A., and Shkrebyonock M.P., "Low Temperature

Expansion Pulse Tubes," Advances in Cryogenic Engineering. Vol. 29, Plenum

Press, New York, 1984, p. 629.

[23] Herrmann , S. and Radebaugh R., "Measurements of the Efficiency and

Refrigeration Power of Pulse—Tube Refrigeration," National Bureau of Standards

Technical Note 1301, September 1986.

[24] Radebaugh, R. and Herrmann, S., "Refrigeration Efficiency of Pulse Tube

Refrigerators," Proceedings of the 4th International Crvocooler Conference. Boulder, Colorado, 1986.

[25] Scott, R.E., "Notes on Linear Circuit Analysis," Northeastern University,

Copyright 1956, Addison—Wesley.

[26] Radebaugh, Ray, Chowdhury, K., Zimmerman, Jim, "Optimization of a Pulse Tube

Refrigerator for a Fixed Compressor Swept Volume," Proceedings of the 5th

International Crvocooler Conference. 1988.

77
APPENDIX A

DERIVATIONS

A.l Derivation of <mc>- Flow Through Annular—Gap Isothermalizer

Assumptions: Ideal gas, laminar flow

dP/dx = f (m/Ag)2/2prh (1)

f= 24/Re = 24/(4rh/M)(m/Ag) = 6///rh(m/Ag) (2)

dP/dx = 3/f (m/Ag)//9rh2 (3)

m = (/9rh2Ag/3//)(dP/dx)

= (rh2Ag/3//RT)P(dP/dx) (4)

Integrate:

<m> = (rh2Ag/3//RT) (Ph2-Pi2)/2L

= (rh2Ag/)/6/^L) [(Ph/Pi) + 1] AP (5)

For Isothermalizer 2:

Th = t/2 t = annular gap thickness

Ag = wt w = annular gap width

Flow Impedance:

Zf = 12L/wt3 (6)

<riic> = (pwt3/24//L) [(Ph/Pi) + 1] AP

<mc> = {p/2) [(Ph/Pi) + 1] (AP///Zf) (7)

78
A.2 Derivation of <mh>: Flow Through Orifice

Assumptions: Ideal gas, adiabatic flow, (AP/P) <<1

<mh> = Am/(r/2) = (2/r)(m2-mi) (1)

mi = PiV/RTi (2)

m2 = P2V/RT2 (3)

T2 = Ti (P2/Pi)(^^^/^ (4)

m2 = (V/RTi) P2 (P2/Pi)^^"^^/'^ (5)

<mh> = (2/r)^ [P2(P2/Pi)^"^/^-Pi] (6)

- (2/r) 1^ [(P2/Pi)^/^-ll (7)

-(2/r)^ [(AP/Pi + 1)1/^-1] (8)

Using a binomial expansion for small AP,

<mh> = (2/r)^ 1 AP

RTTL 7 Pi (9)

<mh> = (2/r)-^|^ (10)

A.3 Mass Flow Rate Through Isothermalizer 2 for Turbulent Flow

Since flow rates with N2 are so high, flow will not be laminar. Therefore, an

expression for AP with the friction factor and the impedance factor Zf must be found.

79
^p = f^^(^Mgp (1)

Re = -^ (m/Ag) (2)

m/Ag = -^

(3)

^^ - ^2/)ih Ag 4rh - 8/? XJF];? ^^^

Z^ - ^

(6)

AP = fRe^^^ (7)

Trial and error procedure:

(1) Guess in (in turbulent region)

(2) Calculate Reynolds No., Re =

4(m/Ag)rh

(3) Find Fanning Friction Factor from Fig. 7-4 in Kays and London

(4) Calculate AP

AP = fRe = ^

Zf = 6.870 X 105 cm-3

//j^ (10 atm, 200 K) = 132 x 10-7 Pa-s

80
p^ (10 atm, 200 K) = .01721 g/cm3

(5) Compare experimental and calculated values of AP

(6) Adjust in accordingly and repeat

A.4 Calculation of Refrigeration Power Ratios for Table 3-1.

Experimental:

Constant (\^p /\):

QHe/QN2 =

jlfl = 2.026

Constant (Pd /Pa ):

QHe/QN2 = 11^ = 1.226

Experimental Corrected:

Constant (\^p /M ):

Pa He VHe

^^^J^^^^^

= (2.026)

ri059]

[l069j

[6.0791

[6.070j

fO.181 1 [0.196 J

= 1.856

81
Constant (Pd /Pa ):

(QHe/QN2) = i*^/^*^ M]^

(rtlh/rtlc)He

^^d 1^^ ^ He

= 0.848

Model:

Constant {Wc? /\ )

QHe/QN2 = 7He/7N2 = j^ = 1.1929

Constant (Pd /Pa ):

QHe/QN2 = 7He/7N2 = f^ = 0-8383

82
APPENDIX B

EXPERIMENTAL DATA

B.l Experimental Data Showing Effect of Heat Transfer on

Gross Refrigeration Power in OPTR

Tc Position of Gross

Temp Probe Refrigeration

(K) (cm from (W)

hot end)

149.9 19.95

150.0 6 19.54

149.9 13 17.83

199.5 32.61

200.0 3 30.36

200.0 6 29.15

200.0 10 28.67

199.9 13 27.19

B.2 Experimental Optimum Phase Angle

The following equation is developed from the phasor diagram in figure (2—4) and

gives 9 in terms of measureable quantities:

Tan e = <^^^l^'> (1)

<mh>(Th 1%)

<dMt/dt> = ^^ uj ' (2)

83
Experimental Data for Calculation of Optimum Phase Angle

Data Set Tc <Klli> <dMt/dt>

No. (K) (g/s) (g/s) (deg)

82.0 0.0817 0.150 40.2

1 79.9 0.1225 0.3240 29.1

95.2 0.1943 0.2452 17.8

122.4 0.1334 0.2746 33.9

2 126.4 0.1772 0.2427 24.8

134.4 0.2119 0.2170 20.2

151.3 0.2688 0.1814 15.2

146.5 0.0803 0.2561 51.2

3 130.8 0.1242 0.2440 34.4

128.0 0.1517 0.2240 26.7

129.3 0.1791 0.2060 21.6

90.4 0.0971 0.3918 44.2

4 71.6 0.1124 0.4066 34.6

78.1 0.1640 0.3084 21.4

98.5 0.0505 0.2752 55.0

5 74.0 0.0869 0.2681 31.3

87.3 0.1404 0.2140 19.5

131.3 0.2140 0.1591 14.6

153.9 0.0999 0.3122 52.0

6 135.3 0.1367 0.3005 38.4

125.4 0.1529 0.2853 32.0

122.4 0.1686 0.2657 27.2

84
B.3 Experimental and Theoretical Data

Tc Freq Pa Pd/Pa Mc Mh/Mc Th/T

(K) (Hz) (kPa) (g/s)

110.0 6.098 1198 0.307 2.579 0.132 2.909

150.0 6.061 1016 0.355 1.731 0.184 2.167

150.0 6.098 1206 0.350 2.344 0.157 2.167

150.0 6.135 779 0.337 1.053 0.227 2.167

150.0 6.061 575 0.262 0.636 0.250 2.167

150.0 6.061 1205 0.348 2.328 0.160 2.167

150.0 6.061 1200 0.351 2.256 0.172 2.167

189.0 6.135 1236 0.387 2.288 0.181 1.799

200.0 6.098 1072 0.388 1.924 0.199 1.700

200.0 6.098 1039 0.330 1.365 0.223 1.700

200.0 6.024 1071 0.201 0.734 0.330 1.700

200.0 6.098 1072 0.323 1.452 0.335 1.700

200.0 6.098 1025 0.292 1.115 0.300 1.700

200.0 6.098 1057 0.321 1.855 0.203 1.700

200.0 6.061 1076 0.388 1.883 0.184 1.700

200.0 6.061 1085 0.388 1.424 0.327 1.700

200.0 6.024 1089 0.325 1.905 0.248 1.700

84.9 5.98 1252 0.290 4.4 0.11 3.651

85.0 6.02 1245 0.301 4.190 0.095 3.647

95.6 5.98 1204 0.307 3.44 0.16 3.295

95.5 6.06 1188 0.319 3.200 0.126 3.298

120.0 6.06 1146 0.346 2.170 0.220 2.667

119.8 6.02 1135 0.364 2.000 0.148 2.671

119.5 6.06 1156 0.336 2.34 0.23 2.678

151.1 5.95 1247 0.367 2.05 0.29 2.151

151.1 5.05 1250 0.379 1.970 0.240 2.151

188.5 5.88 1222 0.385 1.3 0.54 1.804

187.2 5.85 1223 0.396 1.250 0.460 1.790

187.5 5.95 1225 0.410 1.170 0.280 1.787

209.8 5.95 1155 0.397 1.25 0.53 1.668

234.7 6.02 1221 0.420 1.090 0.490 1.534

234.7 6.02 1220 0.445 1.090 0.280 1.532

235.7 6.09 1218 0.409 1.07 0.58 1.527

274.1 6.02 1211 0.421 1.03 0.66 1.368

273.6 6.13 1213 0.433 1.020 0.530 1.371

275.0 6.06 1215 0.440 0.942 0.410 1.364

275.1 6.09 1217 0.460 0.905 0.230 1.363

85
B.3 Experimental and Theoretical Data (Continued)

EXP

Qr/RTc Model 1 Th/Exp Exp Model 2 Model 3 Th/Exp Th/Ex

(J/g) 2<fi>/mcRT Ratio Qgross

(W)

<fl> <fl> Ratio 2 Ratio

(W) (W)

0.0337 0.1852 5.49 9.94 134.86 65.29 13.57 6.57

0.0686 0.2223 3.24 18.51 183.23 74.85 9.90 4.04

0.0546 0.1870 3.43 19.95 184.09 73.64 9.23 3.69

0.0816 0.2604 3.19 13.40 166.54 71.01 12.43 5.30

0.0666 0.2229 3.35 6.60 103.99 45.62 15.76 6.91

0.0551 0.1895 3.44 20.01 185.25 74.26 9.26 3.71

0.0573 0.2055 3.59 20.15 200.02 80.91 9.93 4.02

0.0635 0.1979 3.12 28.54 241.96 83.14 8.48 2.91

0.0852 0.2062 2.42 34.08 229.92 76.39 6.75 2.24

0.0713 0.1965 2.76 20.23 212.39 71.62 10.50 3.54

0.0649 0.1820 2.80 9.9 200.26 75.27 20.23 7.60

0.1080 0.2889 2.67 32.61 322.22 120.50 9.88 3.70

0.0770 0.2339 3.04 17.84 249.42 89.68 13.98 5.03

0.0845 0.1740 2.06 32.60 191.33 63.71 5.87 1.95

0.0703 0.1906 2.71 27.54 212.09 69.88 7.70 2.54

0.0938 0.3388 3.61 27.76 380.08 140.77 13.69 5.07

0.0707 0.2152 3.04 28.00 240.86 82.69 8.60 2.95

0.0297 0.1830 6.17 11.52 108.80 60.18 9.44 5.22

0.0325 0.1638 5.04 12.03 97.63 53.08 8.12 4.41

0.0374 0.2542 6.79 12.80 161.10 88.87 12.59 6.94

0.0433 0.2083 4.81 13.84 131.82 69.32 9.52 5.01

0.0658 0.3189 4.85 17.80 240.81 122.98 13.52 6.91

0.0483 0.2260 4.68 12.03 167.68 77.86 13.94 6.47

0.0578 0.3251 5.62 16.80 246.61 128.78 14.68 7.67

0.0875 0.3596 4.11 28.17 359.72 166.54 12.77 5.91

0.0865 0.3073 3.55 26.78 308.17 132.60 11.51 4.95

0.1253 0.5890 4.70 31.92 680.49 569.65 21.32 17.85

0.1132 0.5121 4.52 27.53 593.68 310.54 21.56 11.28

0.0715 0.3222 4.51 16.30 381.17 142.16 23.38 8.72

0.1249 0.5514 4.42 34.04 658.69 371.90 19.35 10.93

0.1400 0.4959 3.54 37.22 689.09 279.74 18.51 7.52

0.0919 0.2998 3.26 23.13 416.86 128.50 18.02 5.56

0.1591 0.5691 3.58 41.70 801.56 427.01 19.22 10.24

0.1587 0.5971 3.76 46.57 922.75 476.58 19.81 10.23

0.1472 0.4941 3.36 42.70 777.33 274.66 18.20 6.43

0.1059 0.3864 3.65 28.53 605.07 180.11 21.21 6.31

0.0680 0.2265 3.33 17.61 357.20 93.80 20.28 5.33

86
BL-114A

(5-90) U.S. DEPARTMENT OF COMMERCE

NATIONAL INSTITUTE OF STANDARDS AND TECHNOLOGY

BIBLIOGRAPHIC DATASHEET

PUBUCATION OR REPORT NUMBER

NIST TN 1343

2. PERFORMINQ ORQANIZATION REPORT NUMBER

3. PUBUCATION DATE December 1990

4. TITLE AND SUBTITLE Analytical Model for the Refrigeration Power of the Orifice Pulse Tube Refrigerat Lor

5. AUTHOR(S)

Peter J. S torch, Ray Radebaugh, and James E. Zimmerman

6. PERFORMING ORQANIZATION (IF JOINT OR OTHER THAN NIST, SEE INSTRUCTIONS)

U.S. DEPARTMENT OF COMMERCE

NATIONAL INSTITUTE OF STANDARDS AND TECHNOLOGY

BOULDER, COLORADO 80303-3328

7. CONTRACT/QRANT NUMBER

8. TYPE OF REPORT AND PERIOD COVERED

9. SPONSORING ORQANIZATION NAME AND COMPLETE ADDRESS (STREET, CITY, STATE, ZIP) NASA/Ames Research Center

Moffett Field, CA 94035

10. SUPPLEMENTARY NOTES

11. ABSTRACT (A 200-WORD OR LESS FACTUAL SUMMARY OF MOST SIGNIFICANT INFORMATION. IF DOCUMENT INCLUDES A SIGNIFICANT BIBUOGRAPHY OR

UTERATURE SURVEY, MENTION IT HERE.)

The pulse tube refrigerator is a cryocooler with potential for high reliability. This

regenerative refrigerator operates with low average pressures, small compression ratios,

and only one moving part. In research conducted at the National Institute of Standards

and Technology, a pulse tube refrigerator with an orifice located at the closed, hot end

of the system has been used to reach 60 K in one stage. A gross refrigeration power of

12 W has been measured at 80 K with a pulse tube volume of 30 cm'^ . This paper presents

the development of an analytical model to predict the refrigeration power produced in an orifice pulse tube. Three versions of the model are presented: the first calculates

quantities which are independent of dead volume; the second neglects dead volume; and the

third describes a system with dead volume. Experimental results confirm predictions by

the model that refrigeration power is proportional to the average pressure, the pulse

frequency, the mass flow ratio, and the square of the dynamic pressure ratio. Results are presented that are useful for approximate design of orifice pulse tube systems. The model assumes adiabatic conditions and an energy balance is performed in terms of net enthalpy

flow. Phaser analysis was found to be very useful for analyzing the oscillations of

pressure, temperature, and mass flow rate.

12. KEY WORDS (6 TO 12 ENTRIES; ALPHABETICAL ORDER; CAPITAUZE ONLY PROPER NAMES; AND SEPARATE KEY WORDS BY SEMICOLONS)

analytical model; cryocooler; cryogenics; pulse tube refrigerator; refrigeration

power, refrigerator

13. AVAILABIUTY UNUMITED

FOR OFFICIAL DISTRIBUTION. DO NOT RELEASE TO NATIONAL TECHNICAL INFORMATION SERVICE (NTIS). ORDER FROM SUPERINTENDENT OF DOCUMENTS, U.S. GOVERNMENT PRINTING OFFICE, WASHINGTON, DC 20402. ORDER FROM NATIONAL TECHNICAL INFORMATION SERVICE (NTIS), SPRINGFIELD. VA 22161.

14. NUMBER OF PRINTED PAGES

94

IS. PRICE

ELECTRONIC FORM

*U-S. GOVERNMENT PRINTING OFFICE: 1991-574-146/45039



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