Albert J. Juhasz

Glenn Research Center, Cleveland, Ohio

A Mass Computation Model for Lightweight 

Brayton Cycle Regenerator Heat Exchangers

NASA/TM—2010-216799

September 2010

AIAA–2010–7087
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Albert J. Juhasz

Glenn Research Center, Cleveland, Ohio

A Mass Computation Model for Lightweight 

Brayton Cycle Regenerator Heat Exchangers

NASA/TM—2010-216799

September 2010

AIAA–2010–7087

National Aeronautics and

Space Administration

Glenn Research Center

Cleveland, Ohio 44135

Prepared for the

8th International Energy Conversion Engineering Conference (IECEC)

sponsored by the American Institute of Aeronautics and Astronautics

Nashville, Tennessee, July 25–28, 2010
Acknowledgments

This work was supported by the Power and On-Board Propulsion Division, 

Thermal Energy Conversion Systems Branch at NASA Glenn Research Center.

Available from

NASA Center for Aerospace Information

7115 Standard Drive

Hanover, MD 21076–1320

National Technical Information Service

5301 Shawnee Road

Alexandria, VA 22312

Available electronically at http://gltrs.grc.nasa.gov

Level of Review: This material has been technically reviewed by technical management. 

This report is a formal draft or working 

paper, intended to solicit comments and 

ideas from a technical peer group.
NASA/TM—2010-216799 1 

A Mass Computation Model for Lightweight Brayton Cycle 

Regenerator Heat Exchangers

Albert J. Juhasz

National Aeronautics and Space Administration

Glenn Research Center

Cleveland, Ohio 44135 

Abstract

Based on a theoretical analysis of convective heat transfer across large internal surface areas, this 

paper discusses the design implications for generating lightweight gas-gas heat exchanger designs by 

packaging such areas into compact three-dimensional shapes. Allowances are made for hot and cold inlet 

and outlet headers for assembly of completed regenerator (or recuperator) heat exchanger units into 

closed cycle gas turbine flow ducting.

Surface area and resulting volume and mass requirements are computed for a range of heat exchanger 

effectiveness values and internal heat transfer coefficients. Benefit cost curves show the effect of increasing 

heat exchanger effectiveness on Brayton cycle thermodynamic efficiency on the plus side, while also 

illustrating the cost in heat exchanger required surface area, volume, and mass requirements as effectiveness 

is increased. The equations derived for counterflow and crossflow configurations show that as effectiveness 

values approach unity, or 100 percent, the required surface area, and hence heat exchanger volume and mass 

tend toward infinity, since the implication is that heat is transferred at a zero temperature difference. To 

verify the dimensional accuracy of the regenerator mass computational procedure, calculation of a 

regenerator specific mass, that is, heat exchanger weight per unit working fluid mass flow, is performed in 

both English and SI units. Identical numerical values for the specific mass parameter, whether expressed in 

lb/(lb/sec) or kg/(kg/sec), show the dimensional consistency of overall results. 

Nomenclature

a passage height dimension (<0.01 m for plate-fin HX)

AW ≈ (AW + AC)/2 average surface area 

Ch = (Wcp)h hot fluid capacity rate (W/K)

CC = (Wcp)C cold fluid capacity rate (W/K), also Cmin., and Cmax, are, respectively, the smaller and 

larger of the CC and Ch values

CPC ,CPT specific heats for the compressor and turbine flows 

dCAS density of casing material (kg/m3

) 

eC surface effectiveness for cold area, AC

eh surface effectiveness for hot area, Ah 

εRG regenerator effectiveness

hC ,hh local heat transfer coefficients, cold and hot sides (W/m2

-K (SI) or Btu/ft2-hr-R 

(English))

H and T gas enthalpy and temperature values with subscripts referring to state points

k thermal conductivity of the wall material (units in W/m-K (SI) or Btu/ft-hr-R 

(English))

MCAS regenerator casing mass (kg)

tW casing wall thickness (m)

TIC compressor inlet temperature (K)

TIT turbine inlet temperature (K)

TIT /TIC cycle temperature ratio

WC, WT compressor and turbine mass flow rate
NASA/TM—2010-216799 2 

VCOR regenerator core volume (m3 or ft3) 

S total heat transfer surface area (m2 or ft2

) 

U overall heat transfer coefficient (W/m2

-K or Btu/ft2-hr-R) 

Introduction

The function of heat exchangers is to transfer thermal energy between different media, such as a 

warmer fluid, which is being cooled, and a colder one that is being heated in the process. The fluids may 

be either in the liquid or gaseous phase, or one or the other may even undergo a change of phase, as in the 

example of a boiling liquid or a condensing vapor. Heat exchangers are essential components in space 

heating, refrigeration, air conditioning, and chemical and petrochemical plants and refineries. They also 

have a major role in the implementation of various thermodynamic cycles in large electrical power plants 

and heat engines, including Closed Cycle Gas Turbines (CCGT) for stationary conventional fueled or 

nuclear power plants (Refs. 1 to 3) or propulsion systems (Ref. 4) where efficient conversion of thermalto-electrical energy or rotary shaft work energy is a principal objective. For the variety of heat exchanger 

design configurations for each particular application, which may range from shell-tube to plate-fin

designs in parallel flow, counterflow, or crossflow configurations, refer to the many texts in this field

(Refs. 5 and 6). For a given temperature difference between the hot and cold fluids, the underlying 

objective for all heat exchanger types is to maximize the heat transfer surface area of the wall or 

separating barrier between the two fluids, while minimizing the resistance to flow, or pressure drop of the 

fluid flow through the heat exchanger. Since the focus of this paper is on regenerator surface area and 

weight, refer to Reference 7 for an analysis on heat exchanger pressure drop. See Reference 8 for an 

edited compilation on the topic of gas turbine heat transfer. 

The objective of this paper is to develop a theoretical model for the determination of CCGT

regenerator (recuperator) heat exchanger volume and mass for power systems ranging from kilowatt to 

multi-megawatt output levels. The model development includes an examination of the fundamental 

relationships between the required heat exchanger effectiveness, heat transfer surface area and the internal 

heat transfer coefficients available, both on the hot and cold passage side for overall heat transfer as an 

input variable. This input variable can be obtained from the thermodynamics of the Closed Brayton Cycle 

(CBC) power system under consideration as modeled, for example, by the BRMAPS code (Refs. 9 and 

10), the total working fluid mass flow rate passing through the regenerator’s hot and cold passages.

Development of Regenerator Mass Model

As indicated above, modeling of a CBC recuperator heat exchanger needs to be performed in 

conjunction with the analysis of the entire Brayton energy conversion system to determine the working 

fluid mass flow rate and the temperatures at the entrance and exit state points for the hot and cold flow 

passages. Four possible regenerated CBC configurations for space power systems with a nuclear reactor 

heat source for direct or indirect heat input and rejection are shown in Figure 1. For the indirectly heated 

cycle configurations the input thermal energy is supplied by liquid metal cooled reactors (LMCR), while a 

high-temperature gas reactor (HTGR), which heats the inert CCGT gaseous (e.g., He and He-Xe) working 

fluid directly, is used for the directly heated cycles. Similarly, indirect heat rejection implies a pumped 

liquid-to-gas heat exchanger cooling the cycle working fluid, while the liquid is cooled in a space 

radiator. In direct heat rejection the cycle working fluid flows over evaporator ends of radiator heat pipes, 

which in turn the condenser sections are cooled via the space radiator.

Note that removing the recuperator from each of the four configurations shown would result in four 

additional cycles referred to as non-regenerated CBC configuration. 

Since these non-regenerated power systems do not require a recuperator, they are not the main focus 

of this paper. However, they are used in system tradeoff studies in comparing the higher expected cycle 

efficiency of regenerated systems at certain operating conditions with the lower system mass of the nonregenerated CBC configurations.
NASA/TM—2010-216799 3 

CBC Regenerator Heat Transfer

To review the role of the regenerator in a CBC gas turbine power system refer to Figure 2, which 

shows a direct comparison of a regenerated CBC schematic (Fig. 2(a)) with its corresponding 

Temperature-Entropy or T – S diagram (Fig. 2(b)).

The nomenclature of Figure 2 is largely self explanatory with C and T designating “compressor” and 

“turbine,” H.S. is heat source, Q is thermal energy, and WC and WT represent work (for compressor and 

turbine). Temperatures are designated by T and pressures by P, with subscripts referring to the various 

locations in Figure 2(a). 

Figure 1.—Regenerated CBC configurations with direct or indirect heat input and rejection.

Figure 2.—Regenerator effect on Brayton cycle. (a) Schematic diagram. (b) T – S diagram.
NASA/TM—2010-216799 4 

A figure of merit called the regenerator effectiveness, εRG, is the ratio of the actual quantity of heat 

transferred, QTRANS, to the compressor exit stream at temperature TOC to the heat available for transfer in 

the turbine exit flow, QAVAIL at temperature. Hence,

( )

( )

( )

T PT ( ) OCOT

C PC QIN OC

T TOT TOC

C TQIN TOC

AVAIL

TRANS

RG TTCW

TTCW

HHW

HHW

Q

Q

−

− = −− ==ε (1) 

where, by definition 

εRG regenerator effectiveness

WC, WT are the compressor and turbine weight (i.e., mass) minus flow rate

CPC ,CPT specific heats for the compressor and turbine flows

H and T gas enthalpy and temperature values with subscripts referring to state points shown in 

Figures 2(a) and (b) 

From the numerator of Equation (1) it is easily seen that

QTRANS = C PC ( ) QIN −TTCW OC (2a) 

but is also true that

QTRANS = T PT ( ) OT −TTCW REJ (2b) 

As previously mentioned, in closed cycle applications, WC = WT with specific heat being a constant 

for helium or for any inert gas mixture, Equation (1) simplifies to

TOT TOC

TOT TREJ

TOT TOC

QIN TOC

RG TT

TT

TT

TT

−

− = −

− =ε (3) 

To further illustrate how the presence, or absence, of a regenerator (recuperator) affects the system 

schematic and the T – S diagram of a CBC, consider the comparisons shown in Figure 3. 

Note that whereas the regenerated cycle heat is transferred from the higher temperature, and the lower 

pressure turbine exit streams to the compressor flow, there is no heat exchange between the two streams 

for the non-regenerated cycle, which therefore needs to operate at a higher cycle pressure ratio, POC /PIC,

to maximize thermodynamic cycle efficiency. Since at this higher pressure ratio, the turbine exit 

temperature TOT may be approximately the same or even lower than the compressor exit temperature TOC ,

any attempt to transfer heat between the two streams would actually lower the cycle efficiency and 

therefore be counterproductive. So, while the non-regenerated cycle has a simpler configuration and a 

lower mass due to the absence of a heat exchanger and connecting ductwork, the regenerated cycle has a 

higher efficiency at its lower optimum pressure ratio. This will be further illustrated in a later section of 

this paper.
NASA/TM—2010-216799 5 

Figure 3.—Thermodynamic comparison of regenerated and non-regenerated CBC. (a) Regenerated cycle. 

(b) Non-regenerated cycle.

Derivation of Regenerator Total Surface Area

Referring to Figure 2(a) one can define an effective, or log mean temperature difference, ∆tm, as

( ) ( )

( )

( ) OT QIN

REJ OC

REJ OC OT QIN

m

TT

TT

TTTT

t

−

−

−−− =∆

ln

 (4) 

In our derivation it is convenient to expand the terms in parentheses of Equation (4) in terms of the 

temperature changes for the compressor stream (TQIN – TOC) and the turbine discharge flow (TOT – TREJ). 

Rewriting Equation (1) to show the dependence of ∆tm on these temperature differences gives

( ) ( )

( ) ( )

( ) OCOT ( ) QIN OC

OCOT OT REJ

QIN OC OT REJ

m

TTTT

TTTT

TTTT

t

−−−

−−−

−−− =∆

ln

 (5) 

Substituting for (TQIN – TOC) and (TOT – TREJ) from Equation (2) into Equation (5) yields

( )

( ) ( )

( ) ( T OCO TRANS C PC )

OCOT TRANS T PT

TRANS PCC TRANS PTT m

CWQTT

CWQTT

CWQCWQ t

/

/ ln

)/( /

−−

−−

− =∆ (6) 
NASA/TM—2010-216799 6 

The somewhat complicated form of Equation (6) can be greatly simplified by making the following 

substitutions for some of the parameter groups: 

C PC

TRANS

CW

Q=β ; 

PTT

PCC

CW

CW

=χ ; and −=δ TT OCOT (7) 

With the indicated substitution Equation (6) can now be rewritten as 

( )

( )

( )

( ) 



 





β−δ

βχ− +

βχ− =

β−δ

βχ−δ

βχ− =∆

1 ln 1

1

ln

1

tM (8) 

Since the only difference between the compressor flow WC and the turbine flow WT is due to potential 

leakage by ducting compressor flow to gas foil bearings and alternator cooling, the value (1 – χ) is very 

close to zero. Expanding the denominator of Equation (6) into a series and neglecting higher order terms 

leads to 

( ) ( )

β−δ

βχ−  ≈



 





β−δ

βχ− +11 ln 1 (9) 

Substituting the expression obtained in Equation (9) into Equation (8) results in the following expression 

for ∆tm: 

( )

( ) ( ) β−δ≡ β−δβχ−

βχ− ≈∆

/1

1

tm (10) 

Using the symbol definitions from Equation (7), Equation (10) can be written as 

( )

C PC

TRANS m OCOT CW

Q TTt −−≈∆ (11) 

Since the total heat transferred QTRANS is a function of total surface area S, Δtm and an overall heat transfer 

coefficient, UAV, the expression for this coefficient is adapted from the previously cited Reference 5, 

which defines the overall heat transfer coefficients for the hot and cold sides, Uh and UC in terms of hotand cold-side heat transfer coefficients hh and hC can be written as

( ) hwhhh ( ) heAAkAA CChC

a

heU

11 1

+= + (11a)

( ) ( ) CwCCC heAAkAA hhCh

a

heU

11 1

+= + (11b) 

The following symbols remain to be defined for Equation (11): 

a passage height dimension (<0.01 m for plate-fin HX)

hC ,hh local heat transfer coefficients for the cold and hot sides (units in W/m2

-K (SI) or 

Btu/ft2

-hr-R (English)) 

k thermal conductivity of the wall material (units in W/m-K (SI) or Btu/ft-hr-R 

(English))
NASA/TM—2010-216799 7 

eC surface effectiveness for cold area, AC

eh surface effectiveness for hot area, Ah 

AW ≈ (AW + AC)/2 average surface area 

Ch = (Wcp)h hot fluid capacity rate (W/K)

CC = (Wcp)C cold fluid capacity rate (W/K), also Cmin. and Cmax, are, respectively, the smaller and 

larger of the CC and Ch values

Kays and London (Ref. 5) next compute a number of heat transfer units parameter defined as

dAU

CC

UA N

A AV

tu ∫ == 0 min min

1 (11c) 

Letting U = UAV (in W/m2

-K), as defined in Equations (11), represents an overall heat transfer 

coefficient, and S (in m2

) stands for a total required heat transfer surface area. The total heat transferred 

from the turbine exit gas to the compressor discharge stream can be written as

TRANS ** ∆= tUSQ m = ( ) 



 



 −− PCC

TRANS

OCOT CW

Q TTUS 

or 

( ) OCOT

C PC

TRANS TTUS

CW

US Q = − 















1+ (12) 

Solving for QTRANS we obtain 

( )

















+

− =

C PC

OCOT

TRANS

CW

US

TTUS Q

1

(13) 

Rearranging Equation (13) and noting from Equation (2b) that QTRANS can be written as QTRANS =

WTCPT(TOT – TREJ), we may write

( ) ( )

PCC

OCOT

TRANS PTT OT REJ

CWSU

TT TTCWQ 11

+

− = =− (14) 

Since for a regenerator (recuperator) with the same working fluid passing through the hot and cold

passages, WTCPT/WCCPC ≈ 1, Equation (14) may be solved for the regenerator effectiveness as defined in 

Equation (3). Thus

TT USCW

TT

TOT TOC C PC

TOT TREJ

RG 1 /

1

+ = −

− =ε (15)

Equation (15) can then be solved for surface area S as given by Equation (16) 

















ε−

ε = RG

C PC RG

U

CW S

1 (16) 
NASA/TM—2010-216799 8 

While Equation (16) expresses the total surface area required for a regenerator of effectiveness εRG, in 

terms of total mass flow rate WC = WT (identical to the compressor or turbine flow rate) and overall heat 

transfer coefficient U, a form found more useful by the author for actual heat exchanger surface area 

calculations is based on surface area per unit working fluid mass flow, or

















− = RG

PC RG

C U

C

W

S

ε

ε

1 (17) 

Multiplying the result obtained on the right-hand side of Equation (17) by the compressor mass flow 

rate will give the total heat exchanger surface area. Equations (16) and (17) show that the surface area S

will be directly proportional to the throughput mass flow rate, and vary inversely with U, and also 

increase logarithmically with effectiveness εRG, becoming theoretically infinite at 100 percent

effectiveness (i.e., εRG = 1).

Determination of Regenerator Mass

Once the total heat transfer surface area has been determined, calculation of regenerator mass can be 

accomplished in a few steps. One necessary additional input is a surface compactness parameter, which 

expresses the number of square units of surface area that can be packaged, or stacked, into a unit volume 

of multipassage regenerator heat exchanger hardware, while still leaving enough open area between the 

stacked plates for hot and cold gas to flow through. Next, the density of materials used for the heat 

exchanger plate-fin construction—the casing and the connecting ductwork—needs to be determined. 

Manufacturing experience has shown that as much as 1000 ft2 of surface area can be stacked into a cubic 

foot of volume. This packing density or stacking factor (STF in English units) translates into ~3281 m2

/m3

in SI units, referred to as STRFRSI in the regenerator subroutine developed for this paper. The 

regenerator core volume, VCOR, can be written as

COR = * STFSV (18) 

Initially assuming a cubic configuration, a characteristic length dimension is (VCOR)

0.333. Arbitrarily 

setting this value as the height, LH(m), we can set the length LL = 2*LH and the width dimension, LW = 

0.5*LH, the core volume for a parallel-piped form can be written as 

COR = LLLV WLH (m3

) (19) 

The regenerator core mass, MCOR (kg), will be the volume times an average core density, dCOR

(kg/m3

), giving

COR = WLH dLLLM COR (kg) (20) 

The regenerator casing mass is the sum of the masses of the six sides of the parallel-piped 

MCAS = *2 [( ) ( ) ( ) ++ ] WWLLHWH ** dtLLLLLL CAS (21) 

where 

MCAS regenerator casing mass (kg)

tW casing wall thickness (m)

dCAS density of casing material (kg/m3

) 
NASA/TM—2010-216799 9 

Finally, assuming that the mass of the hot and cold inlet and outlet ducts is 25 percent of casing mass, 

that is, MDUCT ~ 0.25MCAS, the total regenerator mass MREG becomes

REG COR CAS ++= MMMM DUCT (22) 

The above equations and procedures were programmed into a subroutine called RGNW of the 

previously mentioned BRMAPS code, which was then run in a standalone mode to generate the results 

discussed in the next section. Note that since regenerator effectiveness, εRG , was an input variable for 

these computations, the procedure to compute εRG for the recommended counterflow or crossflow 

recuperator configurations for desired Ntu values (Ref. 5) was not required for the subroutine. The 

procedure is however available as an option in the author’s main program (BRMAPS).

Regenerator Mass Results

Numerical results of regenerator mass are shown in Table 1, which is an output by one of the author’s 

subroutines on regenerator performance and specific mass, for He working fluid, as a function of 

regenerator effectiveness (ERG), stacking factor (STFRI), core material density (DENI), and overall heat 

transfer coefficient (UU). Resulting values are shown in columns 2 and 5, respectively, referring to 

specific heat transfer area and specific mass. As shown, the first 10 rows of the table are displayed in 

English and the remaining rows in SI units. The units displayed in column 5 are identical to the SI and 

English unit portions of the table.

TABLE 1.—REGENERATOR DESIGN COMPARISON IN ENGLISH AND SI UNITS FOR He WORKING FLUID 

(a) English units

Regenerator 

effectiveness

(ERG)

SPCFC area

(SOW)

(ft2-hr/lb)

Stacking factor

(STFRI) 

(ft2/ft3)

Material density

(DEN=DENI)

(lbm/ft3)

SPCFC mass

(SMASS)

lb/(lb/sec)

Heat transfer coefficient

(UU)

(Btu/(ft2-hr-R)

0.100 0.00551 1000.0 243.5 5.80 25.00

.200 .01240 1000.0 243.5 13.04 25.00

.300 .02125 1000.0 243.5 22.36 25.00

.400 .03306 1000.0 243.5 34.78 25.00

.500 .04959 1000.0 243.5 52.17 25.00

.600 .07439 1000.0 243.5 78.25 25.00

.700 .11572 1000.0 243.5 121.72 25.00

.800 .19837 1000.0 243.5 208.67 25.00

.900 .44633 1000.0 243.5 469.51 25.00

.950 .94226 1000.0 243.5 991.18 25.00

(b) SI units

Regenerator 

effectiveness

(ERG) 

SPCFC area

(SOW)

(m2

-s/kg)

Stacking factor

(STFRSI) 

(m2

/m3) 

Material density

(DENSI)

(kg/m3

) 

SPCFC mass

(SMASS)

kg/(kg/sec)

Heat transfer coefficient

(UUS)

(J/(m2-sec-K)

(W/m2

-K)

0.100 4.06 3280.8 3900.0 5.80 142.08

.200 9.14 3280.8 3900.0 13.04 142.08

.300 15.67 3280.8 3900.0 22.35 142.08

.400 24.38 3280.8 3900.0 34.77 142.08

.500 36.56 3280.8 3900.0 52.16 142.08

.600 54.84 3280.8 3900.0 78.23 142.08

.700 85.31 3280.8 3900.0 121.70 142.08

.800 146.25 3280.8 3900.0 208.62 142.08

.900 329.07 3280.8 3900.0 469.41 142.08

.950 694.70 3280.8 3900.0 990.97 142.08

As a check on the internal dimensional consistency maintained during the calculations, the reader 

may easily verify that with constant overall heat transfer coefficients of 25 Btu/ft2

-hr-R, which is identical 

to 142.08 W/m2

-K. The numerical specific mass values in column 5 are identical within roundoff error 
NASA/TM—2010-216799 10

accuracy in both systems of units. The reason is that the units of lb/(lb/sec) and those of kg/(kg/sec) both 

reduce to units of seconds, so identical numerical values do result. The author has found that the practice 

of performing internal computations in two sets of units is an excellent way to ensure that no dimensional 

errors inadvertently were introduced into computational procedures. The accuracy of code output results 

can thus be validated to a higher degree of confidence level than would otherwise be the case. The SI 

portion of the tabulated results is plotted in Figure 4, for a range of overall heat transfer coefficient values. 

This figure confirms the observations made regarding Equations (16) and (17), concerning the rapid 

increase of total surface area per unit mass flow, especially for regenerator effectiveness values greater 

than 0.8. This is because the surface area tends towards infinity as the effectiveness parameter is 

approaching unity. This is manifested by the heat exchanger surface area and mass doubling as the 

difference between 100 percent effectiveness and the design effectiveness is halved. Thus a recuperator 

having 95 percent effectiveness would have twice the surface area and mass required for 90 percent 

effectiveness. Similarly, an increase in effectiveness from 98 to 99 percent would require doubling the 

surface area and mass of the heat exchanger.

The curve for the low heat transfer coefficient values shown in Figure 4 is representative for 

conventional heat exchangers with multiple staggered tube rows. Space power systems will need 

regenerators utilizing the plate-fin heat exchanger design concept with the capability to raise the heat 

transfer coefficient values to over 280 W/m2

-K (50 Btu/hr-ft2-F). The dramatic reduction in regenerator 

weight for the higher heat transfer coefficient values can be readily deduced from the figure. Furthermore, 

it needs to be pointed out that the regenerator core density value was based on use of high-temperature

metal alloys with values of ~3900 kg/m3 (see Table 1). Future use of high thermal conductivity composite 

materials may reduce this value by a factor of three or more. Such a breakthrough will be of great value in 

further reducing the mass (weight) for heat exchange components of space-based systems. 

Figure 4.—Regenerator specific mass as a function of effectiveness with heat 

transfer coefficient U as a parameter.
NASA/TM—2010-216799 11

Figure 5.—Thermodynamic efficiency of Brayton cycles at various 

regenerator effectiveness values and operating conditions with He 

working fluid. (a) Temperature ratio 3.0. (b) Temperature ratio 4.0.

Comparative Cycle Performance

Having determined the mass penalty on CBC space power systems due to incorporation of a 

recuperator (i.e., regenerative heat exchanger) in the energy conversion cycle, tradeoff studies on 

regenerated versus non-regenerated cycles require relevant information on how regenerators designed for 

a range of effectiveness values will enhance cycle efficiency. Figure 5 shows results of thermodynamic 

cycle efficiency as a function of compressor pressure ratio for He working fluid with regenerator 

effectiveness, ERG, as a parameter. These results were obtained using the author-developed CBC code, 

referred to as BRMAPS (Ref. 9). Cycle performance for cycle temperature ratios of 3.0 and 4.0 are shown 

in Figures 5(a) and (b), respectively.

It should be noted that due to the increased cycle pressure drop of the highly regenerated cycle 

(ERG = 0.95), the turbine pressure ratio was set at only 94 percent of the compressor pressure ratio, 

implying a loss pressure ratio (LPR = 0.94) for a well designed regenerator. For the non-regenerated cycle 

(ERG = 0), the turbine pressure ratio was 98 percent of that developed by the compressor, and the LPR 

parameter was set proportionately between these values for the lower ERG parameters shown. The reader 
NASA/TM—2010-216799 12

may surmise that an excessively high regenerator pressure loss would negate the increase in cycle 

efficiency obtained by regeneration. This is indeed a correct conclusion. If the LPR parameter is set to 

0.8, implying a 20 percent pressure loss, the efficiency for 95 percent effectiveness would be no higher 

than that obtained for the non-regenerated cycles.

Regarding Figure 5, note that cycle efficiency is raised considerably as regenerator effectiveness

(ERG) is varied from 0, indicating no regeneration, to 0.95 (i.e., 95 percent) at pressure ratios below a 

certain threshhold, or crossover value, which is ~3.7 for IC (ALFA) = 3.0 and ~5.4 for ALFA = 4.0. 

Detailed numerical calculations of cycle state point temperatures confirm that the crossover pressure 

ratios designate conditions at which the turbine exit gas temperatures (TOT) are exactly equal to 

temperatures at the compressor discharge, and thus, there is no heat transfer from the turbine discharge 

gas stream to the compressor stream. Hence, all cycles will have the same efficiency, regardless of 

regenerator effectiveness. For pressure ratios above the crossover value, the efficiency of the 

nonregenerated cycle will exceed that of the regenerated cycles with the highest effectiveness regenerator 

incurring the largest efficiency penalty.

Additional observations that can be gleaned from Figure 5 show that the pressure ratio at which the 

highest cycle efficiency is obtained becomes progressively lower with increasing effectiveness (εRG = 

ERG). Thus, for a cycle temperature ratio of 3.0 (Fig. 5 (a)), the optimum pressure ratio is approximately 

1.8 for an εRG of 0.95, yielding an efficiency of about 0.36 (i.e., 36 percent). But for the lower ERG value

of 0.80 the pressure ratio optimum rises to ~2.20, at which an efficiency of only 30 percent is obtained. 

The reader may also note that as the cycle temperature ratio is raised to 4.0, efficiency values increase by 

~10 percent and the optimum pressure ratios shift to the right by increments that are in inverse relation to 

the ERG value. The increase in cycle efficiency between the non-regenerated case (ERG = 0) and the 

highest regenerated cycle (ERG = 0.95) at the optimum pressure ratios for each is about 10 percent for 

each temperature ratio. Note that tradeoff studies of cycle efficiency versus regenerator mass also need to 

take into account that, whereas centrifugal turbomachinery may be sufficient for the relatively low 

pressure ratios required for highly regenerated cycles, the higher optimum pressure ratios for the nonregenerated machines may require multistage axial turbines and compressors. These, however, have a 

higher component (i.e., polytropic) efficiency, which may significantly reduce or even negate the 

efficiency advantage of regenerated cycles. Detailed tradeoff studies will require accurate performance 

turbomachine specifications. 

Concluding Remarks

A computational procedure for the determination of CBC (closed Brayton cycle) recuperator mass 

and an accompanying computer subroutine were developed for incorporation in a previously written gas 

turbine code. Relevant equations were derived for computing the total heat exchanger surface area 

requirements as a function of overall heat transfer coefficient, plate-fin material thermal conductivity, and 

gas turbine mass flow rate. Relevant results on recuperator mass per unit CBC flow rate (in kg/(kg/sec) or 

lb/(lb/sec)) as a function of effectiveness with overall heat transfer coefficient as a parameter were shown

(Fig. 4). As expected, as effectiveness values approach unity, or 100 percent, the required surface area

and hence, heat exchanger mass tends toward infinity since heat transfer at near-zero temperature

difference is implied. The beneficial effect of regeneration on CBC thermodynamic efficiency was also 

shown (Fig. 5), especially for low-cycle pressure ratios, where the temperature of the turbine exit flow 

significantly exceeds that of the compressor. The requirement for low regenerator pressure loss 

(<6 percent) was also implied in the results shown in Figure 5. 

Comparison of Figures 4 and 5 should permit meaningful tradeoff studies on the efficacy of 

regeneration for selected operating conditions. To verify the dimensional accuracy of the regenerator 

mass computational procedure, all calculation steps for a regenerator’s specific mass, that is, heat 

exchanger weight per unit working fluid mass flow are performed in both English and SI units. Identical 

numerical values for the specific mass parameter, whether expressed in lb/(lb/sec) or kg/(kg/sec), show 

the dimensional consistency of overall results. 
NASA/TM—2010-216799 13

References

1. Ackeret J.C.: “Aerodynamische Waermekraftanlage mit geschlossenem Kreislauf,” VDI.85, No. 22, 

pp. 491–500, 1941. 

2. McDonald, C.F.: “The Key Role of Heat Exchangers in Advanced Gas Cooled Reactor Plants,” Heat 

Recovery Systems & CHP, Vol. 14, No. 1, pp. 7–24, 1994 

3. Juhasz, A.J., Rarick, R.A., and Rangarajan, R.: “High Efficiency Nuclear Power Plants using Liquid 

Fluoride Thorium Reactor Technology,” AIAA Paper 2009–4565. Presented at the 7th IECEC, Aug.

2–5, 2009, Denver, CO. 

4. Zucrow, M.J.: “Aircraft and Missile Propulsion-Vol. II,” John Wiley & Sons Inc., New York, 1958.

5. Kays, W.M. and London A.L.: “Compact Heat Exchangers,” McGraw-Hill Book Company, Inc., 

1984. 

6. McAdams, W.H.: “Heat Transmission,” McGraw-Hill Book Company, Inc., 1954. 

7. Sparrow, E.M. and Liu, C.H.: “Heat Transfer, Pressure Drop and Performance Relationship for InLine, Staggered and Continuous Plate Heat Exchangers,” Int. Journal of Heat & Mass Transfer, 

Vol. 22, pp. 1613–1625, 1979. 

8. Sunden, B. and Faghri, M., eds.: “Heat Transfer in Gas Turbines,” MIT Press, Boston, 2001. 

9. Juhasz, A.J.: “Analysis and Numerical Optimization of Gas Turbine Space Power Systems with 

Nuclear Fission Reactor Heat Sources,” Doctoral Dissertation, Cleveland State University, May 25, 

2005. 

10. Juhasz, A.J.: “Multi-Megawatt Gas Turbine Power Systems for Lunar Colonies,” NASA/TM—2006-

214658, Dec. 2006. 
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A Mass Computation Model for Lightweight Brayton Cycle Regenerator Heat Exchangers 

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Juhasz, Albert, J. 

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14. ABSTRACT

Based on a theoretical analysis of convective heat transfer across large internal surface areas, this paper discusses the design implications for 

generating lightweight gas-gas heat exchanger designs by packaging such areas into compact three-dimensional shapes. Allowances are 

made for hot and cold inlet and outlet headers for assembly of completed regenerator (or recuperator) heat exchanger units into closed cycle 

gas turbine flow ducting. Surface area and resulting volume and mass requirements are computed for a range of heat exchanger effectiveness 

values and internal heat transfer coefficients. Benefit cost curves show the effect of increasing heat exchanger effectiveness on Brayton 

cycle thermodynamic efficiency on the plus side, while also illustrating the cost in heat exchanger required surface area, volume, and mass 

requirements as effectiveness is increased. The equations derived for counterflow and crossflow configurations show that as effectiveness 

values approach unity, or 100 percent, the required surface area, and hence heat exchanger volume and mass tend toward infinity, since the 

implication is that heat is transferred at a zero temperature difference. To verify the dimensional accuracy of the regenerator mass 

computational procedure, calculation of a regenerator specific mass, that is, heat exchanger weight per unit working fluid mass flow, is 

performed in both English and SI units. Identical numerical values for the specific mass parameter, whether expressed in lb/(lb/sec) or kg/

(kg/sec), show the dimensional consistency of overall results. 

15. SUBJECT TERMS

Heat exchangers; Regenerators; Recuperators 

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