an analysis of the brayton cycle as a cryogenic ... - nist

jbrary, E-01 Admin. BJdg.

OCT 7 1968

NBS

366

An Analysis of the Brayton Cycle

As a Cryogenic Refrigerator ITATIONAL BURSAL'OF iTANDABD8

LIBRABT

MAR 6 J973

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v S. DEPARTMENT OF COMMERCENational Bureau of Standards

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NBS TECHNICAL NOTE 366

ISSUED AUGUST 1968

AN ANALYSIS OF THE BRAYTON CYCLE AS ACRYOGENIC REFRIGERATOR

R. C. MUHLENHAUPT AND T. R. STROBRIDGE

Cryogenics Division

Institute for Basic Standards

National Bureau of Standards

Boulder, Colorado 80302

NBS Technical Notes are designed to supplement the

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Price $1.25

TABLE OF CONTENTS

PageABSTRACT 1

1. Introduction 1

2. Refrigerants „ . 2

3. Nomenclature List . 2

4. Discussion of Process 4

5. Parameter Values 4

6. Ideal Gas Calculations 8

7. Real Gas Calculations 15

8. Pressure Drop „ 16

9. Results 18

10. Discussion of Results 19

11. Conclusions 30

12. References 33

13. Appendix 34

List of Figures

Figure 1. Schematic Flow Diagram of the Modified BraytonRefrigeration Cycle „ 5

Figure 2. Temperature-Entropy Diagram of the ModifiedBrayton Refrigeration Cycle 6

Figure 3. Temperature-Entropy Diagram of an Ideal BraytonRefrigerator 7

in

TABLE OF CONTENTS (CONTINUED)

PageFigure 4. Schematic Diagram of the Brayton Refrigerator

Performance with an Ideal Gas as the Working Fluid . . 12

Figure 5. Schematic Diagram of the Brayton Refrigerator

Performance Showing a Double Minimum Curve ..... 22

Figure 6. Schematic Diagram of Enthalpy Difference as a

Function of High Pressure For Parahydrogen 24

Figure 7. Schematic Diagram of Brayton Refrigerator

Performance with an Ideal Gas as the WorkingFluid - Single and Double Minimum Curves . 25

Figure 8. Brayton Refrigerator Performance - HeliumRefrigerant - 300 K Cycle Inlet Temperature -

100% Expander Efficiency - No Pressure Drop 39

Figure 9« Brayton Refrigerator Performance - HeliumRefrigerant - 300 K Cycle Inlet Temperature -



70% Expander Efficiency - No Pressure Drop ...... 41


50% Expander Efficiency - No Pressure Drop ...... 42

Figure 12. Brayton Refrigerator Performance - HeliumRefrigerant - 77. 36 K Cycle Inlet Temperature -

100% Expander Efficiency - No Pressure Drop ..... 43





IV


Page


50% Expander Efficiency - No Pressure Drop . . . . . 46


100% Expander Efficiency - No Pressure Drop .... 47







Figure 20. Brayton Refrigerator Performance - ParahydrogenRefrigerant - 300 K Cycle Inlet Temperature -








Figure 24. Brayton Refrigerator Performance - ParahydrogenRefrigerant - 77. 36 K Cycle Inlet Temperature -



Page







Figure 28. Brayton Refrigerator Performance - NitrogenRefrigerant - 300 K Cycle Inlet Temperature -

100% Expander Efficiency - 2 K Heat ExchangerTemperature Difference - No Pressure Drop 59


90% Expander Efficiency - 2 K Heat ExchangerTemperature Difference - No Pressure Drop ..... 60

Figure 30. Brayton Refrigerator Performance - NitrogenRefrigerant - 300 K Cycle Inlet Temperature-70% Expander Efficiency - 2 K Heat ExchangerTemperature Difference - No Pressure Drop 61





VI


Page














50% Expander Efficiency - 6 K Heat ExchangerTemperature Difference - No Pressure Drop ...... 70



vn


Page








70% Expander Efficiency - Pressure Drop. 75


50% Expander Efficiency - Pressure Drop. . „ 76


70% Expander Efficiency - Pressure Drop . 77


50% Expander Efficiency - Pressure Drop 78



Figure 49. Brayton Refrigerator Performance - HeliumRefrigerant - 20.27 K Cycle Inlet Temperature -


vm


Page


70% Expander Efficiency - Pressure Drop. ....... 81








70% Expander Efficiency - 2. K Heat ExchangerTemperature Difference - Pressure Drop 85







Figure 58. Brayton Refrigerator Performance - NitrogenRefrigerant - 300 K Oycle Inlet Temperature -

70% Expander Efficiency - 6. K Heat ExchangerTemperature Difference - Pressure Drop. 89

IX


Page


50% Expander Efficiency - 6.0K Heat ExchangerTemperature Difference - Pressure Drop 90


70% Expander Efficiency - 10.0 K Heat ExchangerTemperature Difference - Pressure Drop 91


50% Expander Efficiency - 10. K Heat ExchangerTemperature Difference - Pressure Drop. ...... 92

Figure 62. Expander Performance - Helium Refrigerant -








Figure 66. Expander Performance - ParahydrogenRefrigerant - 100% Expander Efficiency - NoPressure Drop . . , . 97

Figure 67. Expander Performance - ParahydrogenRefrigerant - 90% Expander Efficiency - NoPressure Drop 98



Page





50% Expander Efficiency - Pressure Drop. 102

Figure 72. Expander Performance - Parahydrogen Refrigerant -



50% Expander Efficiency - Pressure Drop. ...... 104

Figure 74. Specific Flow Rate - Helium Refrigerant 105

Figure 75. Specific Flow Rate - Parahydrogen Refrigerant. . . . 106

XI

AN ANALYSIS OF THE BRAYTON CYCLE AS ACRYOGENIC REFRIGERATOR

R. C. Muhlenhaupt and T. R. Strobridge

The performance of a Brayton-cycle refrigerator has beencomputed taking into account the efficiencies of the various components.The results are presented in graphical form. These charts give the

input power requirements, mass flow rate, pertinent temperatures,and allow selection of the optimum high pressure for a given set of

component characteristics.

Key Words: Brayton, cryogenics, refrigeration

1. Introduction

In 1873 a Boston engineer, George Brayton, conceived an ideal

heat engine which, when operating in a thermodynamic system, combines

constant pressure and constant entropy processes. In actual practice,

the real machine differs from this ideal because of fluid friction in the

flow passages, heat transfer with the surroundings, and irreversibili-

ties in the expansion machine, compressor, and heat exchanger.

In the same year Brayton, using the principles learned in the

development of his heat engine, designed a refrigeration machine which

operated on a reversed Brayton cycle. The working fluid in such a

machine receives heat at some low pressure and temperature and then

rejects heat at some higher pressure and temperature. Brayton'

s

cycle is sometimes referred to as the Joule or the air- standard cycle.

This is because Joule, working independently of Brayton, proposed the

same cycle, and air is commonly used as the working fluid.

During the past several years, the Brayton cycle, operating

as an open cycle refrigerator, has been used extensively in aircraft

cooling, and the cycle is considered by many to be the prime candidate

for space power systems (Bernatowicz, [1963]; Harrach and Caldwell,

[1963]; Glassman and Stewart, [1964]; Schultz and Melber, [1964];

and Glassman et alv [ 1965]). A modified version of this refrigeration

cycle has been analyzed and will be considered here.

Analytical expressions are derived to determine values for flow

rate, high pressure, evaporator inlet temperature, and figure of merit

(the power input per unit of refrigeration) for selected values of ex-

pander efficiency, cycle inlet temperature, evaporator outlet temper-

ature, and cycle temperature difference (the temperature differential

across the warm end of the heat exchanger). The resulting data are

presented in figures 8 thru 75.

2. Refrigerants

Fluids considered in this analysis are nitrogen, parahydrogen,

and helium. The state equations of Strobridge, [1962], Roder and

Goodwin, [l96l], and Mann, [ 1962] were used to calculate the required

thermodynamic properties. Data were obtained for refrigeration at,

and above, the normal boiling points of each fluid which are 77. 364 K

for nitrogen, 20. 269 K for parahydrogen and 4. 214 K for helium.

3. Nomenclature List

C specific heat at constant pressure (J/g mole»K)

C specific heat at constant volume (J/g mole»K)

F volumetric flow rate (scfm)

F/Q compressor flow per unit

refrigeration (scfm/kW)

M molecular weight (He 4.0028 g/g mole)(H2 2.016 g/g mole)(N2 28.016 g/g mole)

P pressure (atm)

R universal gas constant (8. 3143 J/g mole»K)

T temperature ( K)

T ambient temperature (300 K)a

AT temperature difference between1-5

the fluid streams at warm end

of heat exchanger (T - T ) ( K)15W /Q figure of merit (compressor

power per unit of refrigeration) (W/W)

W /F expander power per unit flow (W/ scfm)

h specific enthalpy (J/g)

k specific heat ratio (C /C ) (unities s)p v

m mass flow rate (g/s)

s specific entropy (J/g»K)

scfm standard cubic feet per minute

measured at 273. 15 K and3

1.0 atmosphere (ft /min)

3v specific volume (cm /g)

v , specific volume of the gas atstd p 5

273. 15 K and 1. atmosphere (cm /g)

T\ isothermal compressor

efficiency (unitless)

Q unit of refrigeration (W)

T| isentropic expander efficiency (unitless)

p density (g mole/-t)

101.3278 a multiplication factor (J/^«atm)

30.002119 a multiplication factor (J«scfm/cm «W)

4. Discussion of Process

A schematic flow diagram of the modified Brayton refrigeration

cycle is shown in figure 1, and the process paths are illustrated on tem-

perature-entropy coordinates in figure 2. The thermodynamic processes

involved are isobaric cooling between stations 1 and 2, adiabatic expan-

sion between 2 and 3, isobaric heat absorption of the refrigeration load

between 3 and 4, isobaric heating between 4 and 5, and isothermal com-

pression at ambient temperature between the low and high pressures .

For comparison, figure 3 shows the process paths of Brayton'

s

ideal refrigeration cycle. The main difference between the two figures

occurs in the ideal compression and expansion processes which are isen-

tropic in figure 3 rather than isothermal and adiabatic as shown in

figure 2.

5. Parameter Values

The performance of the refrigerator depends on the fluid used as

the refrigerant, the pressure levels in the fluid circuit, the refriger-

ation temperature, the temperature at the warm inlet of the heat exchanger,

and the efficiencies of the compressor, heat exchanger, and expansion

The process path shown in figure 2 between stations 5 and 1 is notisothermal, indicating that a finite temperature difference will exist inthe heat exchanger of an actual refrigerator. It is assumed that thetemperature of the fluid at the inlet of the compressor will be the same asthe discharge due to heat transfer with the surroundings through the inter-connecting piping.

w,

COMPRESSOR

W.-*/\ EXPANDER'e

EVAPORATORWWWFigure 1. Schematic Flow Diagram of the Modified Brayton

Refrigeration Cycle.

LjJ

(ZZ>H<DCUJ

UJ

ENTROPYFigure 2. Temperature-Entropy Diagram of the Modified

Brayton Refrigeration Cycle.

LU(TZ)H<(ZLUQ_

LU

ENTROPY

Figure 3. Temperature -Entropy Diagram of an Ideal BraytonRefrigerator.

engine. The power required for the compressor was in all instances cal-

culated from the relation for the 100 percent efficient isothermal com-

pression of an ideal gas at 300 K between the appropriate pressure levels.

The results presented here must therefore be corrected to take into

account the isothermal efficiency of actual compressors. The efficiency

of the heat exchanger is reflected by the temperature difference between

the fluid streams at the warm end of the heat exchanger. Table I lists

the values of the parameters used in the analysis.

TABLE I

Parameter Variations

T 300.0, 77. 36, & 20.27 K

T4

8-100 K (helium)

78-100 K (nitrogen)

22-100 K (parahydrogen)

AT 0.5, 1.0, 2.0, 3.0, 4.0, 6. &10.0 K

P 2-60 atm

P c 1 and 1. 2 atm

11 50, 70, 90 & 100%

6. Ideal Gas Calculations

Charts showing the performance of the modified Brayton refriger-

ator assuming that the refrigerants behave as ideal gases are useful in

two ways. First, such charts, and the relations from which they are

derived, are helpful in explaining the characteristics of the various

curves. Second, when the ideal gas performance charts are prepared

as transparent overlays to graphs calculated from real gas properties,

the direct visual comparison shows the necessity for using real gas

properties if accurate results are required.

8

The efficiency (or the figure of merit, as it is commonly called)

of a refrigeration device is usually defined as the ratio of the net power

input to the refrigeration effect. For the purpose of this analysis, the

figure of merit is defined as the ratio of the isothermal compressor

power input (rather than the net power input) to the refrigeration effect.

Thus,

PRT !n(--)W a \P_ /

c 5

Q Tl C (T - T.)c p 4 3

(1)

The compressor power W is negative because it represents power

transferred to the system. However, for convenience, the absolute

value is always used in this analysis.

Since the efficiencies of the expander and the heat exchanger

are important considerations to the refrigerator designer, these proper-

ties must be included. An expression equivalent to (1) is

wc

Pk" X

a VP5

/

Q P -k " 1

^cl^eL 1 -^) j[ T4+AT

l-5J-iT

l-5}5

(2)

For a given set of parameters (i.e., T , AT , T| , and the

refrigerant), there is an optimum high pressure which will require a

minimum amount of compressor power to produce a unit of refrigeration.

This fact can be demonstrated by differentiating (2). Thus

k-1

wQ

d P I

T T] -illa c k {\b-(ir)

k

]T4+AT

1-5J- AT

,,}

1 Tk-1

1

\2

^{\[-(p:)k

]T.+AT,

r_ 4 1 - 5_

AT !-»}'

T 1] T] -^- ( T A+ AT

1a c e k \ 4 1-5/ V P

2k- 1P P

1 \ k("p:Mt£

k-1

IN k

k-1

H P, nc lie 1 - T

A+ AT

! C_ L 4 l-5_

- AT ,-J

which becomes

WQ

d P.

TaV T

4+ AT

l-5 ) {k ~ 1]

Tl P. kc 1

ATk-1

1-5

V T4+AT

l-5>kte) 1+ In

k-1

1 \ k

]}(3)

k-1

i\b-te) T + AT AT \1-5 J 1-5J

For simplification, two terms in the numerator are defined as

follows:

ATA = 1 ,-

1-5

ne(T4+ATl _ 5

)

and

k-1 k-1

B =(p7) '

[— (p1

)

k]

10

For the combinations of component parameters, the term A is

a constant greater than zero but less than one, while B varies from a

maximum of one (as P approaches P ) to a minimum of zero (as P15 1

approaches infinity).

At a pressure P equal to P , the term B equals 1. For this1 5

condition only, (3) becomes

/W

\

rdhf)i T

a(k - 1}

Qd P

T \ kATl-5

Pl

(4)

The terms in (4) are positive; therefore, the slope is negative, and the

curve is decreasing with increasing pressure. Referring to (3), as the

pressure ratio is increased, B decreases causing the slope to increase

until some pressure P w .» is reached where A and B are equal. Here

the slope goes to zero establishing a minimum point on the curve.

Beyond this pressure the difference between A and B is always positive;

therefore, the slope is also positive. As the pressure is increased fur-

ther, the slope, still remaining positive, begins to decrease and continues

to do so until the pressure becomes infinite. Here the slope goes to zero.

Equation (2) was evaluated holding AT. _, T| , 7] , and P_ con-1-5 e . c 5

stant while varying P and T . The results were plotted as a function

of P . The general shape of the curves is shown schematically in

figure 4. For each T at constant AT the optimum high pressure can

be found. For each AT at constant T again the optimum pressure can

be calculated. Therefore, it is possible to plot the locus of optimum

pressures for both constant T „ and constant AT, ...4 1-5

To determine these optimum pressures, (3) is set equal to zero,

and after simplification, the expression becomes

11

uso

+J nnJ •rH

u ^V 14

M•i-i

£<+H

a; > rn

(D Oi

.—

1

a) nJ

^ <v4-> X)<-H

t—

1

a

s0)

1rt

•i-i 0)

Q <J

ooj

-t->

oi a

a<i)m

43

co Ph

^'

<D

U3bfl

0/ 3M

12

AT. . ,P..-^± .P.k" X

I1

-11 (t'+'aT, J-Ip

1) [

1+tn(p:) ]

U °-< 5 >

Tl (T +4T ) ..... -5

The optimum pressure P was then calculated through an iteration

process for several sets of P , T , and AT for the ideal gas refriger-

ants, helium and hydrogen. For comparison to the corresponding real

gas data, these charts are shown as transparent overlays to figures 8

thru 11 and 20 thru 23.

For additional information, a family of evaporator inlet (or

expander outlet) temperature T curves is included on these ideal gas,

optimum operating pressure plots. The efficiency of the expansion

device is defined as

T2" T3

Tl =L *

, (6)e 1 ^ - 1 „.

and the temperature-pressure relationship for an isentropic expansion

is

k-1T P2 , 2 M,

In addition

T \P3 1 3'

T2=T

4+AT

l-5(8)

for an ideal gas. Substituting (8) and (7) into (6) and rear ranging> gives

13

p -i^t3 = <t4

+ati- 5){\[(p:)

- 1 ]" 1 )- (9)

5

where Pnequals P and P equals P . The evaporator inlet tempera-

1 CD J

tures were calculated from (9) using the optimum high pressures deter-

mined in evaluating (5).

As AT approaches zero for all refrigeration temperatures,

the optimum high pressure becomes smaller and is equal to the low

pressure in the limit. Although this is an impossible operating condi-

tion, it is informative to calculate the performance since it is a lower

limit for the power required per unit of refrigeration. Equation (1)

cannot be evaluated explicitly for a high pressure of one atmosphere;

however, it can be changed to the equivalent form

/T4

T InW a \ Tc 3

'

W^ZL-T,,) ' 10)

where T equals T when P equals P . Although (10) also cannot be

evaluated explicitly for T equal to T , it can be solved using L'Hospital's

rule. Thus, in the limit as T approaches T

W Tc a

~Q" =Tl Tl t '

(H)c e 4

Equation (11) was used to determine W /Q at one atmosphere high

pressure for all T values for both the ideal and the real gas cases.

14

:

7. Real Gas Calculations

The analysis that follows assumes no pressure loss in the system;

thus,P_ = P, and P_ = P = P n = one atmosphere. In addition T ,, T]

, and2 13 4 5 r

1 e

T are constant. The compressor is 100 percent efficient with respect

to ideal gas isothermal compression. Steady state conditions are

assumed throughout and elevation and kinetic energy contributions to

the fluid energy are negligible. Application of the first law of thermo-

dynamics to the cycle gives

W RTa^(p;(12)u n

cm <h

4. h

3>

Again the convention of using absolute values of W /Q is adopted. If (12)

is evaluated in the same manner as (2) and the results plotted, the curves

will again exhibit the general behavior shown in figure 4.

The development of (13), used to determine optimum pressures is

given in the Appendix. This expression is similar to (5) and equal to

zero at optimum conditions,

,

-i

101;!

2779 )(p.u.P

'

4 3'\ M / \ 1 P

T] T dh T] Tf / „ e 3 ' \ / 1 \ e 3

x!

i - 1) +i -i

t2/id^^ p

2t2

= . (13)

Equation (13) with its logarithmic term cannot be solved explicitly for P ;

therefore, an iteration process was employed to find the correct pressure

within 0.01 atmosphere.

15

As with the ideal gas plots, a family of evaporator inlet temperature

curves is given to provide additional design information. Since the tem-

peratures at stations 2 and 3' are determined in the course of evaluating

(13), the corresponding enthalpys are known. The enthalpy, and there-

fore the temperature, at station 3 is readily determined from the defini-

tion of efficiency

h - h

e h2- h3-

As stated in the ideal gas calculations, the W /Q values at the one-

atmosphere intercept are determined from (11).

Additional charts show the expander power per unit of volumetric

flow as a function of expander inlet temperature, expander outlet tempera-

ture, and expander inlet pressure. These values were calculated directly

from

W h - h

(15)

for each refrigerant.

F 0.002119 v ,

std

8. Pressure Drop

The discussion thus- far has centered upon a simple, two-pressure

refrigeration system. Since fluid friction occurs in the flow passages,

pressure losses will result. Thus, the pressures P and P (see figureJ. L*

1) will not be equal as assumed in the previous analysis. For the same

reason, the pressures P and P will not equal P . Therefore it is desir-

able to predict what, if any, effect this behavior has upon the parameters

of interest.

16

The expression

Wr <nL~o"JPD

W

Q JNPD

'{^eL 1

k-1

NPD

{T_ 1)

2 e1 -

k-1

PD

(16)

derived from (2) for pressure drop and no pressure drop conditions,

predicts the change in the figure of merit W /Q with reasonable pre-

cision. The subscripts PD and NPD refer to pressure drop and no

pressure drop conditions respectively. A similar expre ssion predict-

ing the corresponding shift in optimum pressure has not been developed.

It is impractical to perform a complete set of pressure drop

calculations because a large number of conditions can be conceived.

Therefore only a brief examination of this phase of the work was per-

formed. Equation (13) was modified slightly to obtain the necessary

optimum operating expression

d

Wf—E-\ Q

L d P1 T

= (h4 " h

3 )+

101. 32779M p

i

tep;

1

1 - T\ +e

H Te 3

'

T~

d h.

d P.

\ T3<

2T2

d P.

d P1 T

= (17)

Pressure drops of 10 and 20 percent in the high and low pressure streams

respectively, believed to be reasonable values were selected for the calcu-

lations. As before, an iteration process was employed to determine the

correct high pressure to within 0. 01 atmosphere.

17

9. Results

The data for optimum pressure plots for helium and parahydrogen

shown in figures 8 thru 27 were calculated using real gas properties.

Several optimum pressure charts are also included for the ideal refrig-

erants helium and parahydrogen. For ease of comparison each ideal gas

plot is a transparent overlay for the real gas refrigerant chart. Each

graph is for a particular expander efficiency and a specific cycle inlet

temperature. The isothermal compressor efficiency in all cases is 100

percent. The figure of merit W /Q is shown as a function of optimum

pressure P for a series of curves representing the evaporator outlet

temperature T , the evaporator inlet temperature T , and the tempera-

ture difference between the fluid streams at the warm end of the heat

exchanger AT1-5

None of the optimum pressure plots show an evaporator outlet

temperature equal to the liquid temperature of the refrigerant. This is

because the optimum pressure is either beyond the range of available data

or too high to be practicable.

Most optimum pressures were beyond the range of available data

for nitrogen; therefore, high pressure--rather than optimum pressure--

plots were obtained for this refrigerant. This information is presented

in figures 28 thru 43. Each chart is for a particular expander efficiency,

a specific heat exchanger temperature difference between the fluid streams

at the warm end of the heat exchanger, and an isothermal compressor

efficiency of 100 percent. The figure of merit W /Q is shown as a func-c

tion of high pressure P for a series of curves representing the evapor-

ator inlet temperature T3

» the evaporator outlet temperature T , the

expander power per unit flow W /F, and the flow per unit of refrigeration

F/Q.

18

Figures 44 thru 61 show performance data for pressure drops

of 10 and 20 percent in the high and low pressure streams of the heat

exchanger. These charts provide essentially the same information and

cover the same general high pressure range as the corresponding ones

for no pressure drop.

Additional performance information for helium and parahydrogen

is given in figures 62 thru 73 which show expander power per unit flow

W /F as a function of expander inlet temperature T , cycle inlete £

pressure P. , and expander outlet temperature T . Figures 62 thru 69

are for zero pressure drop in the heat exchanger, while figures 70 thru

73 are for pressure drops of 10 and 20 percent in the high and low pres-

sure streams respectively of the heat exchanger.

The final two plots (figures 74 and 7 5) show the evaporator inlet

temperature T as a function of evaporator outlet temperature T for a

family of curves representing flow rate per unit refrigeration F/Q.

There are upper and lower limits on the value of T . It can equal (at an

infinite flow rate) but not exceed T . Likewise, it can equal but never

be less than the liquid temperature of the refrigerant (i.e., for the con-

ditions discussed in this work).

10. Discussion of Results

The optimum pressure charts for helium show reasonable agree-

ment with the transparent overlays since the properties of helium,

especially in the high temperature, low pressure regions, are very

similar to those for an ideal gas. The corresponding plots for para-

hydrogen demonstrate the difference between the properties of para-

hydrogen and the ideal gas.

The discontinuities on the 300. K cycle inlet temperature para-

hydrogen curves are in disagreement with the smooth curves on the

19

transparent overlays; therefore, this anomalous behavior must be

examined more closely. Accordingly, a generalized equation

WRTQ

opt

d(P1

)

opt

A + B + C + D_ opt opt opt opt _

M (P,)2

.A 2

1 opt opt

(18)

representing the slope of the constant AT curves may be derived. The

terms in (18) for a variable evaporator outlet temperature T , a constant

cycle inlet temperature T , a constant cycle outlet temperature T , and1- 5

a 100 percent expander efficiency (to simplify the equation) are defined

as follows:

RT /M =a

a constant (approximately 1237),

Aopt

T

opt

d h.

LVdP +

1 T

d h

d P- v.

1 T opt

opt= p

3'

opt

,2 2d h d h

(—,»)+

( tdP, T, d P,

1 1 1

T

d v.

d P1 T opt

C = Popt 1 VT

d T3»

2 ' opt *

d Pl ' T

d h.

[y +

1 T

d h4

- v.d P, ^ 2JIT opt

, and

Dopt

= Pd T.

1 \T J Vd Pn2 opt IT

d h.

d P1 T

d h

d P- v

1 T2 J

opt

It is difficult to obtain precise values for

d h

d P1 T

andd T

d2 h

dP T1

because neither h nor T is an explicit function of P ,IT 5 l

20

and no simple relationship between these properties exists. A brief

examination will be made of the term C which, in fact, is responsible°Pt dh

2dh

4for the discontinuities. The two derivatives {—r——- ) and (

—) are

^ d P, 'rr, ^ d P^rrnIT IT

quite small relative to the specific volume v , thus these three terms

combine to form a negative quantity. The ratio P /T is positive, while

d Tthe derivative ( —

J

is negative (the temperature T decreases1 Tj

with increasing pressure P,). Thus the C term provides a positive

contribution to the slope equation.

Since T decreases as P is increased, there is some pressure

P, ... at which the temperature T„ , reaches the liquid temperature of thel(i) 3 !

refrigerant. When this occurs, the derivative must go to zero. Hence

the positive contribution is no longer available, causing the slope to

change abruptly to some lesser value. The remaining charts do not

exhibit this behavior because the temperature T does not reach the

liquid temperature of the refrigerant for the stated conditions. Had

this happened, a discontinuity would be present.

In a few instances two optimum pressures (i. e. , two minimum

points) were found for parahydrogen precooled to 77. 36 K. The first

minimum generally occurred in the low pressure region of two to ten

atmospheres, while the second occurred in the high pressure region of

90 to 200 atmospheres.. Figure 5 shows a sketch of a typical double

minimum curve which appears to be inconsistent when it is recalled

that the ideal gas calculations predicted only one minimum.

Because the isothermal compressor power W is calculated in

the same way for either the ideal or the real gas, it is evident that this

21

rt di-t u

•Hum

a

Hi H«

>>(1)

0) 1—

1

h 43

PQ

<u Q3

nJ

n &0fl

a "S«$ Oh 43Men<9

Q oa

a

o

a)

co Ph

• I-l

0/ M

22

quantity is not responsible for the unusual behavior. The inconsistencies

therefore must be attributed to the refrigeration load Q which is defined

as

Q = W + m (h_ - h) (19)e 5 1

for both cases. The expander power W changes with increasing pressure

for the ideal gas, while the enthalpy difference remains constant at some

fixed negative value. For the real gas, however, both terms change with

increasing pressure. A typical plot of enthalpy difference as a function

of high pressure for parahydrogen is shown in figure 6. The diagram

shows that Ah is independent of pressure and temperature for the1-5

ideal gas, while it is dependent upon both the high pressure and the cycle

inlet temperature T, for the real gas. The temperature T. is less1 1-a

than T which is less than T ; therefore, the real gas curves begin

to flatten out somewhat and the enthalpy difference tends to approach a

constant value at the higher cycle inlet temperature. Some of the iso-

therms reach a maximum value, at which point the derivative 1 , _VdPl'T

1

must go to zero. For this special condition, and this condition only,

d Tthe Joule-Thomson coefficient f j will also be equal to zero.

1 hl

Therefore, the locus of all such maxima represent the well-known

inversion curve.

If a refrigeration load Q for the ideal gas is recalculated using an

enthalpy difference that varies in a manner similar to that of the T

curve in figure 6, a revised figure of merit W /Q will be obtained.° c rev

Curves representing the original figure of merit W /Q . (solid line)r ° ° ° c orig

and the revised figure of merit W /Q (dashed line) are shown in fig-c rev

ure 7.

23

IT)

i

T|-Q <T|-b<T

l-C

<

deal

rl

Figure 6. Schematic Diagram of Enthalpy Difference as aFunction of High Pressure for Parahydrogen.

24

0/°M

25

«joQ

<L> ID

Pi U* .s° CO

d i

bo -d

h r1

«

o

«O

q•wXUo

CO

CO

u

CO

>

u

"i H 3

e a g

U

fa

At low pressures (point A) the new enthalpy difference is fairly-

constant but slightly greater than the original value. Hence, the revised

curve will lie slightly below the original curve. The minimum (point B)

of the revised curve will occur in approximately the same general pres-

sure region as that of the original, but it will be a lesser value.

As the pressure is increased beyond the minimum point, the new

enthalpy difference becomes somewhat larger. Therefore, the magnitude

of the revised W /Q will be less than that of the original W / Q . ,

c rev c orig

and the W /Q curve should begin to rise (point C). As the pressurec rev

is increased further, the new enthalpy difference reaches its maximum

value causing W /Q to be considerably less than W /Q . . In thisc rev c orig

pressure region (point D) the revised W /Q curve can be expected toc rev

have another minimum or, at the very least, an inflection point. As the

pressure is increased further, the new enthalpy difference begins to de-

crease. In this region (point E) W /Q increases gradually at first,c rev

and then as the magnitude of the new enthalpy difference approaches that

of the original, it increases at a more rapid rate. Thus, for a cycle

inlet temperature less than the maximum inversion temperature of the

refrigerant, the J-T effect will force either a second minimum or an

inflection point to occur on the figure of merit curve. The cycle inlet

temperature of 300.0 K is above the maximum inversion temperature

(about 208.0 K) for parahydrogen; therefore, no extreme deviation from

the normal shape of the W / Q curve should be expected for this condition.

Cycle inlet temperatures of 300. and 77. 36 K are above the

maximum inversion temperature (approximately 55. K) for helium.

Hence no unusual behavior is expected for the curves representing these

conditions. On the other hand, the cycle inlet temperature of 20.27 K

definitely is below the maximum inversion temperature, and although no

double minima were observed for this condition, there was some

26

evidence of an inflection point. This appears to be reasonable because

the maximum deviation of the enthalpy difference curves for helium pre-

cooled to 20. 27 K are considerably less than those for parahydrogen

precooled to 77. 36 K.

For the few cases where double minima did occur, the first was

always selected as the optimum operating condition because it was felt

that the second occurred at a pressure too high to be practicable.

Some of the charts for parahydrogen and all of those for nitrogen

show expansion engine exhaust temperatures that are the fluid saturation

temperatures at the low pressure. This means that a portion of the

refrigerant may change phase from gas to liquid in the expander. In

many units, not designed for this condition, the presence of liquid will

cause a drop in efficiency; however, the operation of "wet" expansion

engines is becoming more common. While the problems encountered

are not insurmountable, it is most important for the designer to be

aware of the situation so that an allowance can be made for the liquid

and for the possibility that a lower efficiency must be tolerated.

The flow data (i.e. , T as a function of T and F/Q) for 1. 2 atmo-

spheres showed close agreement with that for 1.0 atmosphere, with

differences in T of about 0. 01 K and 0. 5 K at the high and low ends of

the T range respectively. Since these differences are negligible and

unreadable on the charts, only the 1.0 atmosphere data are included.

The performance charts may be used in combination to determine

the operating parameters, including the optimum pressure, for a given

refrigeration requirement. Consider the following conditions:

27

Refrigerant

Q

f]

c isothermal

e isentropic

1

helium

100 W60%

70%

300 K

30 K

296 K

0.9 P1

1.2 P,

1.2 P,

1. atm

and

Figure 44 shows that the optimum operating high pressure P is

11 atm for which T and W /Q are 21 K and 32 watts/watt respectively.3 c

The isothermal compressor efficiency must be considered, and this

increases W /Q to 53. 3 watts/watt for the 60 percent efficient compressor.

Thus a 5. 33 kW compressor is needed to provide the required amount of

refrigeration. As indicated, the chart accounts for pressure changes

of 10 and 20 percent in the high and low pressure streams of the heat

exchanger respectively. These losses result in pressures of 9.9 atm

for P_ and 1. 2 atm for P„ and P „.2 3 4

Flow rate information is given in figure 74. For an evaporator

inlet temperature T of 21 K and an evaporator outlet temperature

T of 30 K the flow per unit of refrigeration F/ Q is 250 scfm/kW. Thus

for a refrigeration load Q of 100 watts, the flow is 25 scfm.

The expander performance may be determined from figure 70.

For a high pressure P of 1 1 atm and an evaporator inlet temperature

T of 21 K, the expander output power per unit flow W / F and the

28

expander inlet temperature are 6 watts/ scfm and ~ 35 K respectively.

Thus, for a flow of 25 scfm, the expander output power is 150 watts.

The various refrigeration parameters found from the graphs will

then be

Wc isothermal

WF

Q

e isentropic

35 K

21 K

11 atm

9 . 9 atm

1 . 2 atm

1 . 2 atm

5.33 kW

150 W25 scfm

100 Wand

In order to illustrate the effect of component efficiencies on the

performance of the cycle, consider decreasing AT (i.e., increasing1-5

T ) while holding T| and T. constant. The engine inlet temperature T ,

5 e 1 d.

and thus the evaporator inlet temperature T , will decrease. For the

same refrigeration capability, Q, the mass flow decreases therefore,

the figure of merit W /Q will be smaller. The cycle, however, willc

not be operating at the optimum high pressure which will have a new,

somewhat lower value.

Increasing the isentropic efficiency of the expander T} , while

holding AT and T constant, produces an increase in the expander1-5 1

power output. For the same refrigeration capability, Q, the mass flow

required decreases because of this additional power output and the

value of W /Q becomes smaller. Again the high pressure P can be

reduced to achieve optimum operating conditions.

29

If component efficiencies could be slightly upgraded, considerable

improvement in cycle performance would be obtained. For example,

consider the following design requirements. The refrigerant is helium,

T . is 20 K, T. is 300 K, AT, _ is 6 K, T] is 50 percent, T\ is 60 percent,4 1 1-5 e c

and the pressure drop in the heat exchanger is 10 and 20 percent, res-

pectively, in the high and low pressure streams. Figure 45 gives a

value of 18 5 watts/ watt for W /Q and an optimum pressure of 28 atmc

for a 100 percent efficient isothermal compressor. The 60 percent

efficient compressor increases that value to about 310 watts/watt.

If the component efficiencies could be improved to the following

values

AT1_5

= 4KT] = 70% , and

T) = 70%c

the figure of merit W / Q will be approximately 80 watts/ watt for the

70 percent efficient compressor which is nearly a 75 percent reduction.

In addition, the optimum high pressure will be about 15 atm which

amounts to about a 50% reduction.

11. Conclusions

The proper choice of a low temperature refrigeration cycle to

meet a given set of duty requirements rests on a balance between economic

and technical requirements. Present applications for such refrigerators

are so varied that it is impossible to make definite statements about the

"best" cycle for general use. It is, however, possible to discuss the

advantages and disadvantages of a particular cycle.

30

The modified Brayton refrigeration cycle has basic advantages

which warrant its consideration when choosing a cryogenic refrigerator.

The system is relatively simple, requiring only a compressor, an

expansion engine, and a heat exchanger as its main components. Com-

pared to cycles using Joule- Thomson expansion valves which produce

cooling isothermally only at the liquid temperatures of the refrigerant,

the modified Brayton cycle permits a large selection of refrigeration

temperatures. However, unless specific design allowances have been

made, difficulty could be encountered in attempting to cool at liquid

temperatures with the Brayton cycle because of the gas to liquid phase

change which would occur in the expansion engine.

The optimum high pressures for helium and parahydrogen are not

excessive so the problems of designing expansion engines for large

pressure ratios may be eased. While computing the optimum pressures,

it was evident that the minima in the performance curves were quite

wide, so it would be possible to operate the refrigerator at pressures

lower than the optimum without suffering a severe loss of efficiency.

The Brayton cycle refrigerator is capable of reaching potentially

high thermal efficiencies with reasonable improvements in present heat

exchangers, expanders, and compressors, and competitive performance

can be obtained with present components. However, calculations show

that utilizing reheat expansion stages may increase the available refrig-

eration by 10 to 20 percent with no change in compressor capacity or

input power requirements. Two-stage reheat involves expanding the

high pressure gas not to the low pressure, but to some intermediate

pressure. The cold gas passes through the first portion of the load

exchanger which is now in two sections, where it is warmed to the

expander inlet temperature as it absorbs energy from the refrigeration

31

load. After the final expansion to the low pressure, the fluid enters the

second section of the load heat exchanger. Additional expansion stages

and heat exchanger sections may be added if warranted. In addition to

improved performance, reheat reduces the temperature rise in the load

heat exchanger appreciably and tends to a more uniform heat sink tem-

perature.

Precooling helium to 77. 36 K produces little if any advantage at

the same AT . Precooling to 20.27 K, however, results in a pro-1-5

nounced reduction of the figure of merit at the expense of an increase

in the optimum operating pressure. Precooling parahydrogen to 77. 36 K

definitely decreases both the figure of merit and the optimum operating

pressure. The benefit to be derived by precooling either refrigerant

to 65. rather than 77. 36 K is negligible.

If the heat exchanger efficiency can be improved, both the optimum

operating pressure and the figure of merit will be reduced. In addition,

these values can also be decreased through the use of a more efficient

expander. On the other hand, if a lower refrigeration temperature is

required, the compressor power will be increased.

Any heat leak to the low temperature region of the refrigerator

will decrease the amount of useful refrigeration. Thus the calculated

performance is better than can be realized in practice. The magnitude

of the heat leak depends upon the size and geometry of the low tempera-

ture enclosure and is especially critical in small capacity refrigerators.

32

12. References

Bernatowicz, D. T. (1963), NASA Solar Brayton-cycle Studies, Symposium

on Solar Dynamic Systems, Interagency Advanced Power Group,

"Washington, D. C.

Glassman, A. J., R. P. Krebs, and T. A. Fox (1965), Brayton-cycle

Nuclear Space Power Systems and their Heat Transfer Components,

American Institute of Chemical Engineers 61, No. 57, 306-314.

Glassman, A. J. and W. L. Stewart (1964), Thermodynamic Character-

istics of Brayton- cycles for Space Power, J. Spacecraft 1, No. 1,

25-64.

Harrach, W. G. and R. T. Caldwell ( 1963), System Optimization of

Brayton-cycle Space Power Plants, ASME Winter Annual Meeting,

Philadelphia, Pennsylvania.

Mann, D. B. (1962), The Thermodynamic Properties of Helium from

3 to 300°K between 0. 5 and 100 Atmospheres, NBS Technical

Note No. 154, National Bureau of Standards, Boulder, Colorado.

Roder, H. M. and R. D. Goodwin (1961), Provisional Thermodynamic

Functions for Parahydrogen, NBS Technical Note No. 130,

National Bureau of Standards, Boulder, Colorado.

Schultz, R. L. and W. E. Melber (1964), Brayton-cycle Radioisotope

Space Power System, SAE National Aeronautical Space Engineer-

ing and Manufacturing Meeting, Los Angeles, California.

Strobridge, T. R. (1962), The Thermodynamic Properties of Nitrogen

from 64 to 300°K between 0. 1 and 200 Atmospheres, NBS Tech-

nical Note No. 129, National Bureau of Standards, Boulder,

Colorado.

33

13. Appendix

The development of (13) proceeds in the following manner. A

series of initial cycle calculations were performed using the basic

equation ( 12)

,P

1RT £nW a VPc

Q ^ M(h4

- h3

)

Again the output data were plotted as a function of high pressure for

several values of evaporator outlet temperature, and as expected, these

curves also contained minimum points. Therefore, it was again neces-

sary to derive a suitable means of determining their value.

Differentiating (12) with respect to pressure P , and setting the

derivative equal to zero, yields

(h4 " h

3> ¥1+

dh. P3 1

d P1 5

where P now represents the optimum operating pressure.

Because the enthalpy h is not an explicit function of the pressure

P , the derivative (dh /dP ) can not be evaluated directly; hence, an

equivalent relationship must be found.

Taking an energy balance around the heat exchanger (see figure 1)

gives

h2

= hl

"h5+ h

4 '(21)

34

Differentiating (21) with respect to pressure P yields

d h.

dpiJ

d h.

Ld P.(22)

or

d h2 1

d h.= 1

4 5

The isentropic efficiency of an expander is given by (14) as

(23)

*n=K - K2 3

6 h2" h3-

Solving for h and differentiating with respect to P yields

d h.

LdV = *n

d h

Ld P JIT+ (1 - T1J

d h.

e' Ld P J(24)

Substituting (22) into (24) and (24) into (20) gives

(h4 - h

3>IT+

d h3'

L e Vd P+ (1 - Tl )

1 T

d h.

e' \d P1 T

x tn = . (25)

Because h is not an explicit function of P , the derivative

(dh /dP ) can not be readily obtained; therefore, some relationship

between these two properties remains to be determined. The following

functions are known:

35

hl

= f (pl'

Tl

) {26)

where P is an independent variable, and T is a constant;

h2

= g (hr *V

h5K {27)

where h is a dependent variable, and h „ and h_ are constants:1 .

r 4 5

s2 = j (h

2'P2>- (28)

where h is a dependent variable, and P equals P , an independent

variable; and

h3

,= k(s

2, P

3), (29)

where s is a dependent variable, and P is a constant.

Differentiating the above four equations, yields

dhi= (rp7)„ dp

i• <

30 >

1

d hdh

z= Gnr) dh

i<31 >

1 h4h5

dS2

=(^hT)^ dh

2+ (^) u

dP2 •

and (32)

2 P2

2 h2

d h3»

dh3'

= (^^) pds

2 <33

>

36

respectively. Substituting (30), (31), and (32) into (33) produces the

expression

d h3

=KS s

2 P

d' h. a h1

d h J Vd h J va p2 P. 1 hh 1 T.

2 4 5 1

d P +i \a p

d P2 h.

2J• (34)

Since the pressure at station 2 of figure 1 is the same as that at

station 1, d P must equal d P „ Thus (34) becomes

d h

d P.

a h.

a s2 P

[(£)_(£).. Q) <~ir)2 P_ 1 h^h. 1 T.

2 4 5 12 h.

(35)

From basic thermodynamics it is known that

dh = Tds + vdP,

When the pressure is held constant,

'

9 ^ .T (36)

and when the enthalpy is held constant,

a s \ v3~py,

~h

(37)

37

If the specific volume-density relationship is considered, (37) becomes

5 s

3 P1

PT '(38)

Inserting the proper subindices and substituting (23), (36), and (38) into

(35) yields

d h.3'\ 3'

dVTj

d h1

L\dP1

1 T(39)

1

Finally, substituting (39) into (25), rearranging, and inserting the proper

conversion factors, produces

(h,-h,)+ f";"%?,{, "'

4 3' M A rr n R

[('-.v-%^)11 T d h Tl T_,e 3 ' \ / 1 \ e 3'

d P1 T P

2T2

= , (40)

where the enthalpy h is an explicit function of pressure P , and therefore

the derivative (d h /d P,) T can be readily evaluated.

1

38

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