PB 151397

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cfto. 38

DESIGN AND CONSTRUCTION

OF A LIQUID HYDROGEN TEMPERATURE

REFRIGERATION SYSTEM

BY D. B. CHELTON, J. W. DEAN, AND B. W. BIRMINGHAM

U. S. DEPARTMENT OF COMMERCE

NATIONAL BUREAU OF STANDARDS
THE NATIONAL BUREAU OE STANDARDS

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NATIONAL BUREAU OF STANDARDS

technical <^iote

38

January 12, I960

DESIGN AND CONSTRUCTION OF A LIQUID HYDROGEN

TEMPERATURE REFRIGERATION SYSTEM

by

D. B. Chelton, J. W. Dean and

B. W. Birmingham

NBS Technical Notes are designed to supplement the Bureau's regular publications program. They provide a means for making available scientific data that are of transient or limited interest. Technical Notes may be

listed or referred to in the open literature. They are for sale by the Office of Technical Services, U. S. Department of Commerce, Washington 25, D. C.

DISTRIBUTED BY

UNITED STATES DEPARTMENT OF COMMERCE

OFFICE OF TECHNICAL SERVICES

WASHINGTON 25, D. C.

Price $.75

TABLE OF CONTENTS

INTRODUCTION 2

BASIC CONSIDERATIONS 3

REFRIGERATION SYSTEM 5

REFRIGERATION CONTROL 14

HEAT EXCHANGER CONSIDERATIONS 17

REFRIGERATOR PERFORMANCE 19

APPENDIX - HEAT EXCHANGER DESIGN 27

Design and Construction of a Liquid Hydrogen

Temperature Refrigeration System

by

D. B. Chelton, J. W. Dean and

B. W. Birmingham

ABSTRACT

Maintaining a liquid hydrogen bubble chamber at 27 K has

been achieved with an automatically controlled closed circuit

hydrogen refrigeration system of 300 watts capacity. The system is

sufficiently flexible to be used on other experimental apparatus

requiring refrigeration at liquid hydrogen temperatures. Several

control systems are discussed. Experimental evidence is compared

to predicted performance for design operating conditions.

General design charts are developed that enable heat exchanger

lengths and associated operating parameters to be determined for

the pertinent heat exchanger configuration when employed in liquid

hydrogen refrigerators of other capacities.
2

INTRODUCTION

The use of cryogenics in physics and engineering often requires

continuous refrigeration of apparatus at low temperature *' ' . An

example of this, is the 15 inch diameter liquid hydrogen bubble

chamber at the Lawrence Radiation Laboratory (LRL), Berkeley,

California that requires refrigeration at approximately 27 K. The

Cryogenic Engineering Laboratory of the National Bureau of Standards

has developed an automatically controlled closed circuit hydrogen

refrigeration system with a nominal capacity of 300 watts to meet the

bubble chamber requirements. Previously the bubble chamber

received refrigeration by the evaporation of liquid hydrogen from a reservoir, the liquid being supplied by a remotely located hydrogen

liquefier.

A major design feature of the refrigerator is the use of the

Collins type ribbon packed heat exchanger. '*" The choice of these

commercially available heat exchangers alleviates much of the heat

exchanger design and construction problem. The heat exchanger configuration, because of excessive pressure drop in the low pressure

nitrogen passage, limits the refrigeration capacity to 300-400 watts

for a single three way heat exchanger (see Figure 2). However, the

application of multiple three way heat exchangers effectively increases

the upper limit of refrigeration obtainable.

Design charts were developed that enable the determination of

heat exchanger lengths and associated operating parameters for the

heat exchangers when employed in liquid hydrogen temperature refrigerators. Charts of a general nature are given for hydrogen flow

rates and liquid nitrogen required.

Refrigeration temperatures may be extended to that of liquid

helium by using the hydrogen refrigerator to maintain a precoolant

reservoir in an additional helium circuit . The additional circuit

may be used to produce helium liquefaction or refrigeration to 4. 2 K,

adding considerable flexibility to the apparatus without modification to

the existing hydrogen system.

Manufactured by the Joy Manufacturing Company, Michigan

City, Indiana.
3

BASIC CONSIDERATIONS

The desirable qualities of a closed circuit refrigeration

system can be visualized by considering applicable economics,

logistics and safety. The use of such a system is a more economical

method of obtaining refrigeration than by evaporating liquid from a vented reservoir. The venting of the reservoir (supplied by a remotely located hydrogen liquefier) to the atmosphere results in the

loss of the gas in addition to the loss of the sensible heat of the gas.

In contrast, a refrigerator operates in a closed circuit and utilizes

the sensible heat of the gas obtained from liquid evaporation. The

sensible heat of hydrogen gas between saturation temperature and

room temperature is equal to 8 times the latent heat of vaporization.

The use of the sensible heat reduces the power required to operate a refrigerator to one -third of the power required to operate a liquefier

for equal quantities of refrigeration.

Figure 1 shows the total power requirements of a refrigerator

as compared to a liquefier based on theoretical performance of the

final heat exchanger and 10 K temperature difference at the top of the

H - H - N_ heat exchanger (see Figure 2). The power requirements

have been calculated for a Hampson or Simple Linde cycle with 65 K

liquid nitrogen precooling. The refrigeration obtained from the liquefier is based entirely on evaporation of liquid. The power calculation

assumes 65 percent isothermal compression efficiency and includes

the work necessary to provide precooling. It is assumed that the

nitrogen is supplied to the unit as saturated liquid at 2 atmospheres.

The work to produce the nitrogen refrigeration was taken as 0. 505

Kw-hr per liter of nitrogen ' . Although this may be optimistic for

small scale procurement it serves as a basis for comparison.

For small capacity refrigeration requirements, operating

power considerations may not be of major economic importance, since

the power cost for both a refrigerator and a liquefier will be small in

comparison to other operating costs. However, the power requirement influences the initial capital investment since the cost of a compressor is proportional to the compression power. The compressor for the 300 watt hydrogen refrigerator represents about 30 percent

of the total system investment. The compressor for a liquefier of

comparable refrigeration capacity would be considerably greater. The

cost of purchasing liquid hydrogen as a source of refrigeration is

usually more than predicted by power requirements since the costs

must sustain a completely separate operation for the production and

handling of the liquid. In addition to the increased costs of liquid,

refrigeration logistics require substantial quantities of liquid to be
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5

stored at the use site to insure uninterrupted operation. Losses thus

incurred add to operating expenses.

The handling of a large number of small liquid hydrogen dewars

presents a more formidable safety problem than a closed circuit

system. Each handling increases the possibility of introducing contamination in the containers. Under some circumstances the contaminants may be accumulated in the refrigerated container. In addition, the stored containers required by logistics necessitate a separate area that must be suitably controlled.

REFRIGERATION SYSTEM

The liquid hydrogen bubble chamber at LRL must be maintained at operating temperature on a continuous basis for extended

periods of time (perhaps two or three months. ) Reliability is therefore a prime consideration, and for this reason the Hampson or

Simple Linde cycle was chosen for the refrigerator. Although the

thermodynamic efficiency is not great, the cycle has the advantage of

simplicity and reliability since it avoids moving parts at low temperature.

A schematic arrangement of the refrigeration system is shown

in Figure 2. Saturated hydrogen vapor at 2-10 psig from the evaporator passes to the low pressure side of the final heat exchanger. The

gas is warmed in the final heat exchanger to approximately the temperature of the precoolant fluid by countercurrent exchange with the

high pressure supply gas. The exit temperature is maintained, by

preventing further heat transfer, until it enters the three-way

(H - H - N ) heat exchanger where it is paralleled with the evaporated nitrogen gas from the precoolant bath. By-pass valves are provided on both heat exchangers and transfer line to speed warm-up and

cool-down procedures. The low pressure hydrogen and evaporated

nitrogen in this exchanger is warmed from precoolant temperature to

approximately ambient temperature by heat exchange with the entering

high pressure hydrogen gas. As a result, the high pressure gas is

cooled to slightly above the temperature of the precoolant. The high

pressure hydrogen then exchanges heat with the precoolant liquid in

the liquid-gas exchanger located in the precoolant tank. The tank is

sized to sustain operation for a period of eight to ten hours at design

conditions without additional transfer of liquid. Further cooling of

the high pressure gas is performed in the final heat exchanger. Expansion results in approximately a 50-50 mixture of gas and liquid.

The temperature difference between the high pressure stream and the

low pressure stream at the warm end of the heat exchangers is represented as AT and AT_.
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A small volume momentum chamber is located in the high

pressure line prior to expansion. The purpose of the momentum

chamber is to separate solid impurities that might precipitate out in

the system. The chamber allows solids to collect before they have an

opportunity to obstruct the expansion valve passage. The installation

of such a device has been found desirable in the operation of other low

temperature liquefiers and refrigerators.

The regulators associated with the circulation system shown in

Figure 2 may be classified as feed gas and compressor controls. The

prepurified hydrogen gas is stored in the high pressure (1800 psi) gas

storage cylinders when the refrigerator and evaporator are at ambient

temperature. This gas is drawn into the system through the feed gas

controls as cool-down progresses. Enough gas is stored to fill the

evaporator with liquid to the desired level. Pressure reduction from

the storage cylinder to the compressor intake is done in two steps to

allow the use of a standard commercial unit for removing oxygen

impurity. These controls are valved out of the circuit after the

refrigeration system is charged.

The compressor pressure control valves (PCV-2 and PCV-3)

bypass the compressor when the discharge or intake pressure varies

from their respective set points. The continuous operation of the

compressor tends to reduce the intake pressure when the flow returning from the refrigerator is less than the compressor capacity, i.e.,

less than full refrigeration load. Control PCV-3 bypasses the flow

from the compressor discharge and maintains the intake pressure at a fixed minimum. Control PCV-2 is not normally operative, but

relieves the discharge pressure to the intake when a set point is

exceeded.

High purity hydrogen is used to initially charge the refrigeration system. However, a purification train must be used in the main

process gas stream to remove the impurities (oil and water vapor)

introduced by the operation of the oil lubricated compressor. A micrometallic oil-water separator is located directly after the compressor

aftercooler. Purification is accomplished by the use of an adsorption

unit operating at ambient temperature. The adsorption unit consists

of two cylinders containing 1 3X Linde Molecular Sieves ' " having a relatively high affinity for oil vapor compared to other adsorbents. A

valving arrangement allows flow through the cartridges either in

series or parallel. Although normal operation is in series, parallel

Manufactured by the Linde Company, Tonawanda, New

York.
8

operation permits continuous flow while a contaminated cartridge is

replaced. The unit is designed to allow a minimum contact time of

20 seconds and a maximum superficial flow velocity of 10 feet per

minute under design operating conditions. Design calculation, substantiated by initial operations, indicate cartridge operation will

exceed 5-7 days without regeneration.

The evaporator of the refrigeration system shown in Figure 2

is part of the bubble chamber apparatus. No liquid hydrogen container

is incorporated within the refrigerator proper. However, hydrogen

circulation through the refrigerator may be accomplished independent

of the evaporator by the use of the transfer valves. This arrangement

is desirable when separate cool-down of the refrigerator is necessary.

A more detailed view of the evaporator may be seen in Figure 3. The

liquid hydrogen flask represents the evaporator of the closed circuit

hydrogen refrigerator. The normal refrigeration load is produced by

both the dynamics of the bubble chamber operation and the heat leak

into the low temperature apparatus. The essential parts of the apparatus are surrounded by a liquid nitrogen temperature thermal

radiation shield. The bubble chamber is initially filled by condensation of hydrogen gas within the coils located in the flask. For a given

refrigeration requirement, the liquid hydrogen in the evaporator is

maintained at a constant temperature. Temperature control of the

bubble chamber is accomplished by adjusting the length of the conduction path between the chamber and the refrigerant. Adjustment is

performed with the piston and control rod arrangement shown. Withdrawing the control rod (and piston) increases the distance between the

chamber body and the refrigerant, thus reducing the heat transfer to

the chamber.

Since the refrigeration system was designed to be incorporated

with an existing evaporator, limited space was available for interconnections. The refrigerant and the gas from the evaporator are

transferred through a high vacuum insulated two passage concentric

transfer line. The essential parts of the transfer line assembly are

shown in Figure 4. The vacuum jacket of the transfer line is a copper

tube 1.375 inch outside diameter. The concentric arrangement consolidates the thermal heat influx by allowing only one hydrogen temperature tube to receive room temperature radiation. This reduces

the overall static thermal heat influx to the unit -especially important

where the length of the transfer line becomes appreciable. The

thermal heat load into the transfer line is approximately 0. 4 watts per

foot of length.
LIGHT SOURCE

Figure 3. Cross Section View of 15— Inch

Liquid Hydrogen Bubble Chamber
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Division of the flow on the refrigerator side of the transfer

line is accomplished by a combination of a close tolerance and a metal -to -metal contact, as shown in the figure. An alternate procedure provides a polytetrafluoroethylene "O" ring on the inner liquid

line to replace the metal -to-metal contact. The elastic seal compensates for surface irregularities. Slight leakage of liquid across

the seal does not appreciably reduce the performance of the unit since

the refrigeration of the liquid is utilized in the final heat exchanger to

provide additional cooling of the high pressure gas. Excessive leakage, however, will reduce the efficiency of the heat exchanger to a point that counteracts this compensation. Calculations indicate that

for a design AT =0.25 K (65 K precooling), a 10 percent liquid

leakage reduces the available refrigeration at the evaporator less

than 1 percent. For a design AT =1.0 K,a 10 percent liquid leakage reduces the refrigeration l.n percent.

The evaporator side of the transfer line is provided with a

phosphor bronze braided sleeve to reduce entrainment losses. The

application of the metallic sleeve disperses the entering stream and

allows the liquid to collect in the evaporator in a more quiescent state.

This was found necessary since the evaporator receives a mixture of

gas and liquid. "With the fixed evaporator design it was found impractical to condense the entire stream in a submerged type heat

exchanger. The application of the sleeve increased the refrigerator

performance approximately 15 percent.

Figure 5 shows the entire refrigeration system during a performance test. The major component arrangement is similar to that

shown in the schematic of Figure 2. The oil -lubricated compressor,

capable of delivering 35 scfm of hydrogen at 1800 psi, is located at

the left. A major portion of the circulation system is located within

the panelled framework next to the compressor. The compressor

controls, the feed gas controls, and the purification system are

included. The high pressure gas storage cylinders are located behind

the framework. Manual valves and gauges necessary for operations

performed with the circulation system are on the attached panel. The

second panelled framework, located at the right, contains the dewar

with the refrigerator heat exchangers and the precoolant tank. Refrigerator gauges and manual valves are placed on this panel. An angle

view of the refrigerator framework with the dewar removed is shown

in Figure 6. The transfer line connects the refrigerator to the evaporator located at the far right. The evaporator shown, constructed for

performance testing, consists of a 20 liter liquid hydrogen dewar with

an electric heating element to simulate the refrigeration load normally

provided by the bubble chamber.
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Figure 6. View of Refrigerator Panel Board
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An overall view of the bubble chamber in operation prior to

the refrigerator construction is shown in Figure 7. The refrigerator

is placed adjacent to the bubble chamber. However, the congested

test area requires that the circulation system be placed some distance

away. A small portion of the Bevatron magnet is visible in the upper

left corner. A nuclear radiation shield of heavy concrete separates

the Bevatron from the experimental area. The particle beam is

extracted from the accelerator and is directed through appropriate

passages in the concrete shielding into the bubble chamber. The

concrete bunker in the background normally encloses experimental

apparatus, such as the bubble chamber, that is effected by stray

nuclear fields. These stray fields reduce the clarity of event photographs and add to the difficulty of analysis.

REFRIGERATION CONTROL

Two basic parameters may be varied to control the quantity of

refrigeration at the evaporator. From an energy balance application

of the First Law of The rmodynamics, the controlling parameters are

found to be the mass flow rate and the enthalpy difference at the top of

the final heat exchanger. This may be expressed as follows:

Q = mAH

The variation of these basic parameters presents many possibilities

of refrigeration control. The mass flow rate is affected by the pressure and aperture size of the expansion valve, while the enthalpy

difference of the final heat exchanger is affected by pressure (high and

low), precooling temperature and heat exchanger efficiency. Of these

variables and combinations thereof, three methods of refrigeration

control were selected that could appropriately achieve the desired

results. Control by these methods has been considered by others

and has here been adopted for our particular requirements. '»°

The quantity of refrigeration may be controlled by adjusting

the flow rate through the evaporator while maintaining a constant high

pressure at the expansion valve. The control is performed by positioning the expansion valve in accordance with refrigeration demands.

The position of the expansion valve regulates the required hydrogen

flow while any excess flow from the compressor is by-passed to the

compressor intake by PCV-3. The control signal must be proportional to the change in the refrigeration required and the process

variable must be returned to an initial condition by the corrective

action of the expansion valve. Stable operation requires other system
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variables be held constant and appropriate measures have been taken

to control these variables. For example, a constant precoolant

temperature is assured by a pressure regulator on the precoolant

tank.

The first method for refrigeration control maintains a constant

high pressure by monitoring the pressure prior to the expansion valve

(high pressure control). A variation in the required refrigeration

will affect the compressor intake pressure. Since PCV-3 maintains

the intake pressure above a fixed minimum, a reduced refrigeration

requirement will cause PCV-3 to bypass gas from the compressor

discharge as previously indicated. Consequently, the discharge

pressure tends to be reduced. The pressure reduction tendency is

indicated by an air signal from the pressure transmitter. The expansion valve controller receives the signal and closes the expansion

valve proportionally. The latter action returns the discharge pressure

to the reference point and reduces the hydrogen flow delivered to the

refrigerator. An increased refrigeration demand, not to exceed the

refrigeration capacity of the system, will produce a similar sequence

of events and result in an increased hydrogen flow through the refrigerator while maintaining a constant compressor discharge pressure.

The second method for refrigeration control maintains a constant high pressure by monitoring the level of a low pressure gas

holder (gas holder control). The gas holder control method requires

slight modifications to the schematic shown in Figure 2. Control

PCV-3 may be omitted and the ballast volume defined as a piston

type gas holder. The use of a gas holder is advantageous for several

reasons. The application of a gas holder introduces a time delay

between an operational error and uncontrolled venting. The gas

holder may also be used for a gas storage volume in refrigeration

systems having small evaporator liquid capacities and be used to

reduce compressor surge problems. A variation in the refrigeration

demand results in a gas holder level displacement. The displacement

is translated into the control signal for the expansion valve control,

adjusting the refrigeration hydrogen flow rate as in the high pressure

control system previously described. The expansion valve adjustment

returns the gas holder level to the reference point and thus maintains

constant operating pressures.

A third method of controlling the quantity of refrigeration

regulates both the flow rate and the high pressure before expansion by

maintaining a fixed expansion area (constant aperture control).

Constant aperture control is possible with the arrangement shown

schematically in Figure 2. The expansion valve is manually set to
17

maintain the design operating pressure (1800 psi) at a full refrigeration demand. Further expansion valve adjustment is not required. A

decrease in refrigeration demand tends to lower the compressor

intake pressure, but since this pressure is to be kept constant, high

pressure flow is bypassed. The division of flow results in less hydrogen to the refrigerator and a reduction in the discharge pressure since

the expansion valve has a constant aperture. However, the reduction

in pressure is not excessive since it is not proportional to the reduction in refrigeration demand. In fact, operation of the unit with the

constant aperture control has indicated a discharge pressure reduction from 1800 psi to 1000 psi for a change in refrigeration demand

from 300 watts to approximately 35 watts. An advantage of the

system is that an expansion valve controller is no longer required.

However, an additional pressure regulator is necessary to maintain

design operating pressure in the adsorbers to prevent excessive

velocity and a reduction in contact time.

HEAT EXCHANGER CONSIDERATIONS

The use of the Collins type heat exchanger tube establishes the

gas phase heat exchanger configuration. Therefore, the design

resolves into determining the required heat exchanger lengths and the

associated pressure drops. The liquid-gas heat exchanger located in

the precoolant tank consists simply of a coil of copper tubing. The

heat transfer and pressure drop correlations of Trumpler and Dodge

were used in the design of the gas phase heat exchanger. These

correlations were considered applicable since the present heat

exchanger tubes were geometrically similar. The heat exchanger

design procedures are presented in the appendix.

The effect of the temperature difference of the two fluids

flowing throughout the length of the heat exchanger must be considered

in the design. The location within an exchanger where the temperature of the inlet and return streams have a minimum temperature

difference may be defined as a "pinch" location. This pinch condition

usually occurs at the warm end of a counter -flow heat exchanger using

fluids with constant specific heats. However, a variation of the

specific heat with temperature may cause the pinch location to occur

within the heat exchanger. This internal pinch condition occurs in

some instances in counter-flow exchangers using hydrogen gas.

The pinch condition may be illustrated by plotting the temperature in the exchanger as a function of the total heat transferred. Such

plots are usually called cooling curves. A typical cooling curve is

shown in Figure 8. It is apparent by inspection that the log mean
18

TOTAL HEAT TRANSFERRED

Figure 8. Typical Cooling Curve with Internal Pinch
19

temperature difference (LMTD), for use in the conventional heat

transfer equations, cannot be applied over the entire length of the heat

exchanger by considering only end point conditions. The LMTD can

only be correctly applied by performing a number of stepwise calculations. The number of steps must be determined by examining the

shape of the cooling curves and subdividing it into a series of straight

sections. Average thermal properties of the fluids are used in each

section.

Heat exchanger efficiency has a pronounced effect on the performance of a refrigeration system. Theoretical refrigeration, based

on zero degree temperature difference at the warm end of the final

exchanger, is a useful reference in the initial selection of the characteristics of a given system. Figure 9 indicates the theoretical refrigeration available. However, calculations indicate the quantity of

refrigeration is reduced because of finite temperature differences by

5 percent per degree K with 65 K precooling and 6.7 percent per

degree K for 78 K precooling.

The nominal size of 300 watts selected for the refrigerator was

based on the dynamic load of the bubble chamber with sufficient

additional capacity to facilitate the cool -down of the system and to

condense liquid into the bubble chamber. The quantity of refrigeration

resulted in the choice of a standard compressor having a capacity of

35 scfm (1 atm, 20 C) of hydrogen at 1800 psig, with a precoolant

temperature of 65 K and reasonable heat exchanger efficiency. This

flow rate results in a theoretical refrigeration capacity of about 260

watts when the precoolant bath is operated at atmospheric conditions

(see Figure 9).

Heat exchanger efficiency also influences the liquid nitrogen

required as precoolant for a given system. Figures 10 and 11 show

the liquid nitrogen consumption for various temperature differences

at the warm end of the H - H - N

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heat exchanger. It is assumed

that the nitrogen is supplied to the unit as saturated liquid at 2

atmospheres. Finite temperature differences increase the precoolant

consumption by 4-5 percent per degree K.

REFRIGERATOR PERFORMANCE

The actual refrigerator performance was determined for a number of operating conditions in an effort to verify the method of

calculation. A summary of experimental evidence for the design

operating condition is compared to design predictions in Table I. The

refrigeration efficiency found in the experiments, considering only

the region below the precoolant, indicates 97.4 percent of the design
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Figure 10. Liquid N Consumption for 65°K Precooled Refrigerator
22

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70

Figure 11. Liquid N Consumption for 78°K Precooled Refrigerator

2
23

TABLE I REFRIGERATOR PERFORMANCE

Flow - scfm

Operating pressure - psig

Precoolant temperature - K

Total refrigeration - watts

Refrigeration efficiency

Percent of design

Percent of theoretical

Liquid nitrogen consumption - liters /hr

Temperature differences

(warm end) - K

H

?

- H - N exchanger 14 14

Final exchanger 0.3 0.8

Low pressure KL pressure drop - psi

H

?

- H - N exchanger 5.2 4.0

Final exchanger 0. 55 0.51

Design Experimental

35 35

1800 1800

65 65

338 329

-- 97.4

98. 4 96.0

8. 1 8.3
24

prediction. The total refrigeration produced by the refrigerator must

take into consideration the static heat influx to the system. This, of

course, reduces the refrigeration available at the evaporator.

Experiments indicate a static heat influx of 35 watts. This heat leak

includes the static load on the evaporator and the 14 foot transfer

line with appropriate fittings and connections. The large portion of

the static heat leak is attributed to the evaporator used for the performance test since no effort was made to optimize the cryogenic

design.

The comparisons shown in Table I indicate good agreement

between design prediction and experimental results and justifies the

design procedure used. The performance of the critical final heat

exchanger results in a heat exchanger efficiency, based on temperatures, of 98. 3 percent. This heat exchanger efficiency results in a favorable refrigerator performance.

The design equation used to predict pressure drop within the

heat exchangers is known to be valid within ± 20 percent for a specified

range of Reynolds number. The pressure drop comparisons shown in

Table Tare within the stated tolerance. However, the nitrogen flow in

the H - H - N heat exchanger is insufficient to create a Reynolds

number in the turbulent region (Re< 550). As would therefore be expected, the pressure drop predictions in the nitrogen passage were

found to be considerably below experimental results. In fact, at

design flow rate, the pressure drop was found to be 2. 5 times the

pressure drop predicted by the design equation. The ratio of experimental pressure drop to predicted pressure drop increases at lower

flow rates.

The increased nitrogen pressure drop must be considered in

the attainment of the appropriate precooling temperatures at the

desired flow rate. Refrigerators having hydrogen flow rates below

35 scfm are not materially affected by the increased nitrogen pressure drop since appropriate precooling temperatures can still be

obtained. The design flow of 35 scfm appears to be the threshold for

the attainment of 65 K in the precoolant tank using a single three way

heat exchanger. Flow rates above this threshold require division of

the low pressure nitrogen gas into multiple three way heat exchangers.

The design charts for heat exchanger determination may be used by

selecting the proper heat exchanger flow rate. For example, a total

flow of 60 scfm would be divided into two parallel heat exchangers of

30 scfm each. The heat exchanger characteristics would therefore be

selected from the design charts at 30 scfm.
25

The thermodynamic properties of hydrogen presented by

Woolley, Scott, and Brickwedde A have been used in the design of

the refrigerator and in the determination of the refrigerator performance.
26

BIBLIOGRAPHY

1. R. L. Blumberg, J. D. Gow and A. J. Schwemin "The

Development and Operation of the Ten-Inch Liquid Hydrogen

Bubble Chamber," Proceedings of 1956 Cryogenic Engineering

Conference, K. D. Timmerhaus, Editor, University of

Colorado, Boulder, Colorado, Paper H-l pp. 318-324 (1956)

2. B. W. Birmingham, D. B. Chelton, D. B. Mann, and H. P.

Hernandez, "Cryogenic Engineering of Hydrogen Bubble

Chambers, " ASTM Bulletin, No. 240, pp. 34-39(1959)

3. B. W. Birmingham, "A Mechanical Refrigeration Process for

the No-Loss Storage of Liquid Hydrogen, " Refrigerating

Engineering. Vol. 65, No. 7, pp. 42-44, 86 (1957)

4. D. B. Chelton, J. W. Dean and T. R. Strobridge (to be

published)

5. D. B. Chelton, J. Macinko and J. Dean, "Hydrogen Liquefaction by a Dual Pressure Process, " Refrigerating Engineering, Vol. 65, No. 8, pp. 39-41, 62 (1957)

6. D. B. Chelton, J. Macinko and J. W. Dean, "Methods of

Hydrogen Liquefaction, " NBS Report 5520 (1957) (Unpublished)

7. A. G. Prodell and D. C. Colley, Columbia University, New

York (private communication)

8. G. J. Eckman, University of California, Lawrence Radiation

Laboratory, Berkeley, California (private communication)

9. P. R. Trumpler and B. F. Dodge, "The Design of RibbonPacked Exchangers for Low Temperature Air Separation

Plants, " Transactions of the American Institute Chemical

Engineers, Vol. 42, No. 2, pp. 75-84, (1947)

10. W. H. McAdams, "Heat Transmission," McGraw Hill, New

York, Second Edition, pp. 232, (1942)

11. H. W. Woolley, R. B. Scott, and F. G. Brickwedde

'Compilation of the Thermal Properties of Hydrogen in Its Various Isotopic and Ortho-Para Modifications, " Journal of

Research National Bureau of Standards, 41_, 379 (1948)

RP 1932
27

APPENDIX - HEAT EXCHANGER DESIGN

The physical properties of the Collins type heat exchanger

tubes used in the calculations of the gas phase heat exchangers are

presented in Figure 12. The gas passages (tube annuli) are packed

with a spiral wound ribbon to introduce a large number of fin surfaces

in the gas streams. These fins are soldered to the tube walls to

assure good thermal contact. The active surface area of each gas

passage is the total fin and tube wall surface. The free cross sectional area of a passage is equivalent to the annular area less the area

obstructed by the fins. The tabulated values in Figure 12 were calculated from the tube geometry.

9 Trumpler and Dodge have determined the heat transfer correlation for the Collins type heat exchanger tubes. The standard

dimensionless relationship for heat transfer in fluids flowing without

phase change in turbulent motion through heated or cooled tubes is

used.

Nu = m (Re)

n

(Pr) 1/3

where Nu = Nusselt number

Re = Reynolds number

Pr =• Prandtl number

The coefficient, m, was determined to be 0. 118 and the exponent, n,

to be 0.7. Experimental evidence shows that these values are

applicable for Reynolds numbers ranging from 550 to 6000.

The heat transfer coefficients determined by the solution of

the heat transfer correlation is subject to modification by the application of a fin efficiency term. The fin efficiency was calculated as described by McAdams

tanh ax

N

ax

where x = fin length

a = KS

h, calculated film coefficient

b = exposed perimeter of fin
28

l.500"0.D.

I.084"0.D.

.667"0.D.^Nr .l25"O.D.rod

.086"wall

.025"wall

.025"wall

1.450 I. D.

.034"l.D.

.495"l.D.

Thickness of Ribbon Wall = .018

Width of Ribbon = .052

ANNULUS

NO. of FINS

PER TURN

TOTAL ACTIVE

SURFACE AREA

ft 2 /ft(b)

NO. of TURNS

PER INCH

FREE AREA of

CROSS SECTION

ft 2 (S) ft

I 22 .66 5 .00090 .00548

n 42 1.4 5 .00277 .00792

IE 64 2.1 5 .00402 .00765

Figure 12. Heat Exchanger Geometry
29

S = cross sectional area of fin

K = thermal conductivity

The result of these calculations is presented in Figure 13. Modified

heat transfer coefficients were used at each chosen temperature level

for the calculation of the heat exchanger length. A summary of the

results of these calculations are graphically represented in Figures

14,15, 16 and 17.

The heat exchanger pressure drop was calculated with the aid

of the correlation presented by Trumpler and Dodge.

0. 005185f G

2

AP _ s_

L pDe'

2 G = mass velocity, - #/hr - ft

3

p = density - #/ft

De' = equivalent diameter -in

f = Fanning friction factor

AP = pressure drop -psi

L - length -ft

The density has been evaluated at the arithmetic mean temperature

determined from the heat exchanger end point conditions. The

Fanning friction factor was evaluated from the equation"

f = 0.476 (Re)

~°' Z

These equations may also be applied for Reynolds numbers ranging

from 550 to 6000 as in the heat transfer equations. The calculated

pressure drop is represented in Figures 18 and 19.

The pressure drop for the nitrogen passage is omitted since it

was found that the Reynolds number of the evaporated nitrogen gas was

lower than the recommended value. Therefore the pressure drop

equation was not applicable. A method of estimating the pressure

drop in the nitrogen passages is given in the section on Refrigerator

Performance. The resulting pressure drop is high enough to limit

the attainment of low precooling temperatures by pressure reduction.

This situation may be remedied by a rearrangement of the heat

exchanger above the precooling temperature level. The high pressure
30

flow may be divided into two equal streams and directed through two

parallel three way heat exchangers. After passing through the heat

exchangers the flow would be rejoined before entering the gas -liquid

heat exchanger in the precoolant tank. The multiple heat exchangers

result in reducing the nitrogen pressure drop, making the attainment

of a low precoolant temperature possible. However, the flow in the

heat exchangers must be monitored and adjusted accordingly. A

hydrogen flow rate of 35 scfm proved to be the threshold at which the

attainment of a 65 K precoolant temperature is possible. Flow rates

greater than 35 scfm require the use of multiple three-way heat

exchangers.
31

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33

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34

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35

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36

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37

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2 2 2 2

I S. DEPARTMENT OF COMMERCE

Frederick H. Mueller, Secretary NATIONAL BUREAU OF STANDARDS

A. V. Aslin, Director THE NATIONAL BUREAU OF STANDARDS

The scope of activities of the National Bureau of Standards at its major laboratories in Washington, IXC. and

Boulder, Colorado, is suggested in the following listing of the divisions and sections engaged in technical work.

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Cryogenic Engineering. Cryogenic Equipment. Cryogenic Processes. Properties of Materials. Gas Liquefaction. Radio Propagation Physics. Upper Atmosphere Research. Ionospheric Research. Regular Propagation

Services. Sun-Earth Relationships. VII F Research. Radio Warning Services. Airglow and Aurora. Radio Astronomy and Arctic Propagation.

Radio Propagation Engineering. Data Reduction Instrumentation. Modulation Research. Radio Noise. Tropospheric Measurements. Tropospheric Analysis. Propagation Obstacles Engineering. Radio-Meteorology. Lower Atmosphere Physics.

Radio Standards. High Frequency Electrical Standards. Radio Broadcast Service. High Frequency impedance Standards. Electronic Calibration Center. Microwave Physics. Microwave Circuit Standards.

Radio Communication and Systems. Low Frequency and Very Low Frequenc) Research, High Fre- quency and Very High Frequency Research. Ultralligh Frequency and Super High Frequency Research. Modulation Research. Antenna Research. Navigation Systems. Systems Vnalvsis. Field Operations.