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Comparative Evaluation of Three 3 + 

Altirtnative Power Cycles for Waste z 

Heat Recovery from the Exhaust r' Y 

of Adiabatic Diesel Engines 

M. Mtirray Bailey 

National Aeronautics and Space Aciministration 

Lewis Research Center 

U.S. Department of Energy 

Conservation and Renewable Energy 

Office of Vehicle and Engine R&D 
Comparative Evaluation of Three 

Alternative Power Cycles for Waste 

Heat Recovery from the Exhaust 

of Adiabatic Diesel Engines 

M. Murray Bailey 

National Aeronautics and Space Administration 

Lewis Research Center 

Cleveland, Ohio 441 35 

July 1985 

Work performed fcr 

U.S. DEPARTMENT OF ENERGY 

Conservation and Renewable Energy 

Office of Vehicle and Engine R&D 

Washirrg!~~, D.C. 20545 

Under Intsragency Agreement DE-A101-80CS50194 

a 

5, 

i 1 

DOEINASAI50194-43 i 

NASA TM-86953 ! j 

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1 
Three alternative power cydes were compared in 

application as an exhaust-gas very wern for 

use aith advanced "adiabatic" did eagines. The power 

cyde ahernatives considered were steam Rankine, 

cqmic Rankine with RC-1 as ttir working fluid, and 

variations of an air Brayton cyde. The comparison was 

made in terms of the fuel economy and economic 

payback pad for heavydltty uucks ophg in 

line-had service. The common bascline for the 

comparisons was the performance and cost of an 

adiabaticmmpormd diesel engine in the same 

service. 

"fhe results indicate that, in terms of engine rated 

specific fuel consumption, a did/ahanarivepoawcycle engine offers a significant improvement ova the 

turbcompound diesel baseline. The maximum 

improvement (12 percent) resSed from use of a Rankine 

cycle heat-recovery system in series with 

turhmpourtding. Somewhat less improvement (9 

paom) dted from useof the Rankinecyclesystem m 

nikxhtion for the turbommpounding. 

Performance of the steam Rankine and the RC-1 

organic Ranhe systans were found to be potenh'ally 

equal at the dorninal67 1 'C (1 240 OF) diesel exhaust gas 

tempaature considered in the srudy. The air Brayton 

cyde alternatives studied, which included both simplecycle and compression-intercooled configurations, were 

less effective and provided only about half the fid 

consumption improvement of the Rankine cyde 

alternatives under the same cundhions. 

Highway average fud emnomy was assumed to scale 

dxdy with the identified changes in engine rated 

sWc fuel consumption. The amount of vehicle fuel 

saved with each improvement was then estimated on the 

basis of 161 000 km (100 000 mi) annual use. The value 

of the fuel saved was calculated using the 1W average 

fuel price of 90.32fliter @l.Wgal). 

Caprtal and maintenance cost enimates were developed 

for each of the heat-raovery power cycle systems. These 

costs were integrated with the fuel savings results to 

identify the time required for net annual savings to pay 

back the initial apital investment. An earlier survey of 

industry sourees has indicated thar capital payback 

within a marrimurn of 3 years is required for a con- to 

be coed economically attractive. 

The capaal payback time resuhs were oonsistent with 

the results on improvement in engine rated spesific fuel 

consumption: The Rankine qde heat-recovery system in 

series with tu.rbocompounW showed the kt paykack 

time (3.2 yr), followed by the Rankine cyde system used 

in substitlaion for t~mpormding (4.7 yr); payback 

time for the air Brayton cyde alternatives were 

signifiwtly longer. 

The sensitivity of caphl payback time to arbitrary 

increases in fuel price, not accompanied by 

corresponding hardware cost inflatinn. was examined. 

The results indicated that fuel price in~eases of 5 and 3 

percent, respectively, would be required for the 

turhmpoundplus-Rankine cycle (series) system and 

the Rankinesubstitad-for-tu~kmpound system to 

pay back capital within the maximum acceptable 3 yr 

time. 

Introduction 

The turbcdarged && engine is currently the 

univasaIly wed powaplam for heavyduty, longhaul truck applications. Although considered efficient by 

most standards, these engines nevertheless rejea much of 

their fuel energy in the form of waste heat. The two 

major heat-lm mechanisms are conduction to the water 

jacket and hot gas flowing out through the exhaust stack. 

The adiabatic engine of the future will feature 

insulated cylinders and thus eliminate the water jacket 

mhg. The typical adiabatic configuration will include 

a turbooompound power turbine to reoovet some of the 

waste heat from the high-temperature exhaust. More 

efficient exhaust-heat recovery is desirable, but can only 

be accomplished by the use of a more complex system 

involving an appropriate heat-recovery heat exchanger 

for transferring exhaust heat to a separate power cycle. 

Historically, the DeparMent of Energy (DOE) has 

mvstigated the Organic Rankine Cycle System (ORCS) 

for improved heat recovery from the exhaust of 

conventional did. The advent of the adiabatic engine 

with an exhaust temperature up to 50 permt higher than 

mventional engines suggests a diffaent om 

regime for which the DOE/NASA are currently 

investigating altanative power cydes for exhaust-heat 

raovery. 
This report presents P comparative evaluation of three DLsd Palomme Md FSmud Cortdftlons j .: 

alternative power cycles: steam Rankine, organic 

Rankhe, and air Brayton. The comparison is based on 

cyde data generated in thnt padd NASA study 

contracts plus data generated at the Lewis Research 

Carter on an Automotive-Gas-Turbine (AGT) derivative 

air Brayton system. 

The mmparative evaluation indudes the powa cycle's 

performance, annual fuel savings, cbst, and economic 

payback when used for heat recovery from the exhaust of 

an adiaha& diesel opaaring in typical long-haul truck 

UY. 

This work is part of the Department of Energy WE) 

Heavy-Duty Transport Technolcrgy Program whh projed 

management provided by the NASA Lewis Raearcb 

Center. The specific project dement is Advanced 

Adiabatic Digel Technology; the task is Waste-Heat 

Utihation. 

The power cycle data used in this comparative 

evaluation is based on information developed under 

DOE-funded contracts issued by NASA. The appropriate 

NASA Comraaor Reports (CR's) are irtentifiPn in the 

ref- sedion of this report. 

Adiabatic Diesel Baseline 

The term "adiaMc diesel" as used in this report 

refers to a low heat rejdon engine incorporating 

ceramic mmpanents as required to allow the elimination 

of the traditional wer jacket/radia!or cooling system. 

The engine is potentially more efficient and reliable, but 

it is actually not adiabatic according to the true meaning 

of the term. Current projections indicate that slightly 

ova 30 percent (overall) of the fuel injm energy will be 

rejected in the form of waste heat. 

Waste heat leaves the adiabatic engine primarily in the 

fm of hot gases flowing out through the exhaust stack. 

Significantly, the stack gas edts at temperatures m exass 

of 530 'C (Im OF); making the gas stream an attradve 

heat sotuce for use with most power cycles. 

For purposes of comparing alternative power cycles, it 

was considered desirable to stablish a set of standard or 

baseline adiabatic diesel mnditions. This is particularly 

appropriate for this study, which was conduaed on a 

noninterference basis; that is, the diesel cycle was 

adjusted as it normally would be for best diesel 

The adiaka.tic diesel is currently in an early 

development stage with production of a fdy insulated 

engine not antiapated until at least the mid 1990's. Tht. 

major devdoprnent effort in this country has been the 

U.S. Army Tank Automotive Command (TACOMY 

Cummins Engine Company coopedve program (ref. 1). 

That program has included single-cylinder and 

mutticylinder engine tests with various deges of 

insulation installed. 

The diael data used m this rcport were obtained from 

a series of Dieel Cyde Simulator (DCS) calmhiom. 

(Information provided by V. Sudhakar, C 

Engme Company.) The DCS was programmed to 

gtimate the performance of an insulated engine with no 

water cooling and only minimal heat rewon to the 

engine oil. In this simulation the mcylrnder heat loss is 

reduced 

adiabatic". 

by 60 pacent. Hence the tam "- 

As illustrated in figures 1 and 2, four configurations of 

the 6U-jxrcent adiabatic engine were examined: 

1. T- (TC). 

2. Turbocharged-Afteramled (TC/A). 

3. ~~-~urbocom~ound (TCPD). 

4. Turbocharged-Tuhmmpwnd-Afta~ooled 

CTCPD/AI. 

Th= pararnei& noted on the figures are the rated 

power (full throttle) conditions achieved in each case at 

199 radlsa: (1900 rpm) engk: speed. The simulations 

we established on the basis of an approximately 15 

percent torque rise characteristic typical of engines for 

long-haul applications. The air-to-fuel ratio at rated 

power conditions for each illuarated engine is 28. 

Significant with regard to ptential waste-heat 

recovery are the exhaust-gas temperature and flow rate 

for each engine configuration. Gas temperature is seen to 

vary by a total of 100 'C or 18 percent among the four 

configurarions shown in figures 1 and 2. The highest 

exhaust temperature (671 'C) is associated with the 

turbocharged engine without aftercooling (TC); the 

lowest temperature (571 'C) is associated with the 

tur-mpound engine with aftamling (TCPDIA). It 

should be noted, however, that the general trend of 

specific fuel consumption (sfc) shows higher exhaust-gas 

temperatures associated with h@er relative fuel 

consumption. 

performance. No special compromises were introduced 

to provide increased exhaust energy or otherwise benefit Annd Foel Erpens Correlntfm 

any anticipated altanative power cyde. The resulting Engine performance is related to annual fuel expense 

diesel exhaust conditions were then considered as a frrst through the vehicle charaaeristics that influence 

common heat soura for all alternative power cycle highway fuel economy, then through the amount of 

evaluations. annual ure or driving, and fdy through the average 
Alr inlet 

Ugsd Uhr) 

Alr Inlet 12P5 Uhr E856Mhr) 

(a) Nona£tamM 0. 

(a) - WA). 

prico paid for fud. In terms of relative complexity, thse 

variables may be best discussed m rewrse of the order 

mentioned above. 

Fwd plioe.-Figure 3 illustrates the price range of 

diesel fuel over a 10-year period. The price shown is a 

pump price indudmg average state tax. Bulk or contract 

prices may be Iowa. The data source (ref. 2) is lrpdated 

to indude prices through 1983. 

The fuel price data shows a sharp rise in the 

1978-tck1981 time period fonowed by a modaate dedine 

in 1982 and again in 1983. For purposes of this study, the 

kdbe diesel fuel price is established at the 1983 average ' I. ( price of $0.32/lilcr ($l.22/gal). 

ANUrPl L#.-Annual use refers to the average number 

of highway kilometers (miles) driven per vehide per year. 

A! this point it should be noted thar wasteheat utilization 

systems are targeted primarily for the class 8 traaor- 1 trailer units with gross weights approaching the 36 000 Lg 

(80 000 Ib) limit and operating in line-haul or highway 1 service betwoe. cities. Typical annual use for thae 1 vehicle is considered to be in the range of 161 000 km 

(a) NanaftmM WD). 

(b) Afkadd CTmD/Ak 

(100 000 mi). The transcontinental operators running 

coast-tocoast may average 240 000 km (150 000 mi) per 

year. For purposes of this study, the 161 000 km (100 000 

mi) use rate will be sdoptcd as illmve of the majority 

of long-haul use. 

V&& jid economy.-The most complex dement in 

the annual fuel equation is the determination of 
average mission fuel economy; that is, the actual kmfliter 

(&/gal) experience of the trucks on the road. Ova-theroad or mission average fuel economy is a complex 

intepdon of the engine spe~%c fuel consumption 

characteristics with numerous vehicie and route 

variables. The propa evaluation of all thse variables can 

only be accomplished via a axnputerizbd Vehide-MissionSimulation (YMS) m& (ref. 3). Unfonunatdy, VMS 

modelmg was not available within the soope of this study. 

?he alternative approach used in this study involved 

evaluation of available truck performance data to 

establish trends that could be used to prdct the impact 

of engine sfc improven.ents on mission fuel emnomy. 

For mission fuel emnomy purpcss, two types of 

vehicles were ickdfied; the average truck and the fuel 

saver truck. Charaaeristics of bath truck configurations 

are summarized in table I. The avemge wnfiguration is 

considered to repma! tte median or average of vehicles 

in savice taiay. Portions of the data on this 

configuration wee taken from a current starktical 

abstract of the industry (ref. 4). 

The fuel saver configuration represents the truck of the 

future now bemg developed by the industry (refs. 5 and 

6). It incorporates the latest in state-of-theart equipment 

for fuel emnomy improvement. The engine used is a 

current, waterumled, production unit. The 3.4 kmlliter 

(8.0 milgal) performance reprents 50 percent better 

fuel economy than the avenge truck with only an 11 

penxnt better engine. Obviously, the majority of the fuel 

economy improvement is due to the vehide changes. 

The question of interest for the future is "How will a 

fuel saver confion of this type rerpond to further 

improvements in engine performance?" The precise 

answer could be best estimated with a proper W 

program considering the wmplete map of exh engine 

together with the appropriate vehicle characteristics. A 

less rigorous approach is to assume that, with vehicle 

characteristics constant, the vehide fud economy will 

improve in direct proportion to the change in engine 

rated sfc. Tie validity of this approach is supported by 

studies that indicate that €3 mt of truck mission fuel 

is comumed at throttle setthgs within 10 percent of fullthrottle or rated conditions (ref. 7). 

The curve in figure 4 shows how the fuel economy of 

the fud saver truck would change if engine rated sfc 

improved over that of the engine presently installed. The 

r.. .g 

8 ; - 3 .a :Ii - I m 

- a 

TABLE I.--CHARACTERISTICS OF VEHICLE CONFIGURATIONS 

FOR <SLAES 8 TRUCKS 
curve is based on the assumption that fuel economy will 

change in proponion to the rated sfc change. Lndicated 

on the curve is the 0.178 kg/kWh (0.293 lbhp-hr) sfc 

lwei of the TCPD/A engine of figure 2; this is the best of 

the adiabatic diesel baseline engines. Ewe sfc values 

beiow this Iwd can only be achieved with the addition of 

the heat-recovery power cycler which are the subjed of 

this comparative evaluation. 

Oaerrg condudon.-The final step in the prm is to 

combine the sfc/fuel economy data of figure 4 with 

161 000 km (100 000 mi) annual use to indicate the 

relationship of engine performance to annual fuel usage. 

This overall correlation is illumed in figure 5. The 

annual fud usage of the TCPD/A diesel, shown in figure 

5, is the best diesel performance available, and is thus the 

kdine for measuring improvements available through 

use of the alternative power cycles. 

The objectives of the overall comparison effort 

included an evaluation of economic as well as technical or 

thermodynamic factors. Thus, each conceived dieselplus-alternative-powercycle compound engine system 

must be evaluated not only for the traditional sfc 

performance, but also in tenns of its estimated user cost 

or purchase price as well. This requires the inclusion of 

engine price admates as part of the adiabatic diesel 

baseline data. 

As noted earlier, the adiabatic diesel is in an early 

development stage. Accordingly, purchafe price data are 

nos available. It is expected, however, that the cast of an 

adiabatic diesel will be approximately equal to that of its 

cooled counterpart (ref. 1). The reawning here is that the 

f Refffena hdr with 

vaterzmled diesel 

@-Qnslne--------P 

24 

- 

B 28 33 12 3&lfi2 .. 

2-a j Speclflc W msun;tian. Ltihp-hr 

:el 

.... ! 

,&$ Fm 5.--Codon of ewe spdfic fuel consumpica to annual 

fuel usage for the furl saver vuck cnn!ii. c 

< 4 

elimination of a>olmg system components (radiator, fan, 

pump, and lines) will produce a cost savings that 

approximately offsets the cost of the required advanced 

technology ceramic insulating materials. Amrdingly, 

the price of current production engines is a guide to the 

price of the advanced adiabatic diesel. 

Price data for podudion TC/A engines was gathered 

from several major engine manufacturers and plotted 

together in an attempt to identify a consistent price 

algorithm. The price used was the fleet price, which was 

understood to be approximately 30 percent below the full 

retad or list price. Fleet price is typically associated with 

purchases of 10 or more units; however, r-t depressed 

sales in the trucking industry have resulted in fleet prices 

bang extended to a wider range of buyers. 

The TC/A price data showed some scatter which is 

considered to be typical of price data. After reviewing the 

data, a price of $14 500 was established for the 239 kW 

(320 hp) engine. ho, the data indimted that the change 

in engine price with power level followed a nominal 0.7 

exponent or logarithmic relationship. This relationship is 

illusvated in figure 6 by the solid line identified as TC/A 

diesel. 

With the TC// engine price algorithm established, it 

was next nemsary to dwdop compatible algorithms for 

the other diesel configurations of fqures 1 and 2. This 

was done by assigning a $2000 price premium for 

turbocompounding (including the attendant power gain) 

and a SSOO credit or discount for nonaftermled 

configurations. The resulting engine prices are illustrated 

by the small circles in figure 6. The appropriate dashed 

lines the^ establish the 0.7 exgonent algorithm assumed 

tor estimating the price impact of other engine size levels. 
Maintenance and repair costs include all the labor and 

material expenses required to keep the diesel engine in 

operating condition. For long-haul operations of 161 000 

km (100 000 mi) pa year, these costs are very significant. 

Contrect rep&.-For purposes of this study, diesel 

maintenance and repair costs were estimated on the basis 

of a contract repair program (ref. 8) which provides 

Lifetime average at a levelized annual cost. The coverage 

includes all aspects of maintenance and repair except 

items resulting from neglect or accidents. As such, it 

includes overhauls as required during the contract period. 

A cost breakdown is available which allows separation of 

the conventional engine radiator and fan costs from the 

bulk of the maintenance and repair costs. The data 

available indicated that radiator and fan costs together 

account for approximately 11 permt of the overall 

maintenance and repair cost. 

Fwe 7 illustrates thc adiabatic diesel contract 

maintenance and repair fast correlation developed for 

use in this study. The solid line was developed from two 

cost quota for a conventional, watercooled, TC/A 

diesel. The quotes were obtained for the 224 kW a~d 261 

kW sizes. In each case the conventio~zi engine radiator 

and fan costs were subtraded from the quote to simulate 

the adiabatic engine. Accordingly, the 239 kW (320 hp) 

point on the solid line is the estimate applicable to the 

adiabatic TC/A engine of figure 1. The line follows a 

1.14 exponential relationship. 

The appropriate points in figure 7 represent 

maintenance and repair estimates for the TC, TCPD, and 

Fie 7.-Died maimenana and repair cost mmlations. Levdid 

annual mrts arc based on a 7-year maintenance and repair con!ract 

and 161 000 km (100 000 mi) annual w. 

TCPD/A adiabatic diesels 01 figures 1 and 2. Ln this 

regard, it was assumed that removing the aircooled 

aftercooler with blower (TC/A to TC) would reduce 

maintenance by $200 per year. Discussions with industry 

representatives resulted in an estimated additive cost of 

$330 for maintenance of turbocompounding (TC/A to 

TCPDIA). The dashed lines follow a 1.14 exponential 

relationship allowing extrapolation of costs to oother 

engine power levels. 

Heat Recovery Power Cycle 

Configurations 

The exhaust gws leaving an adiabatic diesel are 

significant in terms of potential recovery power; in fact, 

the potential power of the exhaust stream is nearly equal 

to the shaft output of these already efficient engines. 

Using 149 "C (300 OF) as a reasonable minimum heatrecovay heat exchanger temperature and 0.270 Kcal/kg "C) 

(0.270 FttuAb OF) as the average gas specific heat at 

constant pressure (Cp), the exhaust power potential of the 

TC diesel of figure 1 is 207 kW (271 hp) or 87 percent of 

the shaft power indicated. 

Obviously there is a tremendous potential for wasteheat recovery ad, therefore, improvement in overall sfc. 

The key is to employ a heat-recovery system that 

produces an adequate recovery benefit without exceeding 

tolerable limits on system complexity and cost. 

The numb and variation of heat-recovery power 

cycles and power cycle hardware configurations are very 

large. Significant among the power cycle alternatives 

would be steam Rankine, organic Rankine, and air 

Brayton. These alternatives were the subject of rmtly 

completed conceptual design studies which are the basis 

for this comparative evaluation. Tile Stirling cycle is a 

fourth alternative for which an equivalent conceptual 

design study is underway with a projected completion 

date in 1985. 

Steam Rankine System 

The technology base for steam Rankine systems is, of 

course, eldensive in the areas of electric power 

generation, railway locomotives, and auxiliary power 

generation using pr- steam. The specific background 

for automotive systems dates to the early steam cars. A 

more recent, and probably more significant, base results 

from the Government-sponsored Rankine engine 

development work of the early 1970's (ref. 9). The 

conceptual design and performance data used in this 

comparative evaluation were developed in 1983 in a 

DOE-sponsored study of the steam Rankine power cycle 

for waste-heat recovery from adiabatic diesel vuck 

engines (ref. 10). 
OWQWAL PA=-6 

OF POOR QUALITY 

Rehwt cycle.-The typical high-performance steam 

Rankine system as used in luge electric generating 

stations uses a reheat cycle. Reheat refers to a second 

heating of the partially expanded stm prior to furthe1 

expansion in a low-pressure turbine. The reheat cycle 

provides improved performance but at the expense of 

added heatexchanger surface and expander (turbine) 

hardware. 

The poten;ial of a reheat cycle was explored 

t.armidy for the adiabatic dieseIheat-recovery 

Gpliwion. The analysis indicated that the power cycle 

efficimq wc.uld indeed increase, but that the heatraokerj eiiicitncj, and thus the amount of energy input 

to the cycic, would actually decrease huse the reheat 

element interrupts the true countercurrent operating 

mode of the heat-recovery heat exchanger. The net result 

is a marginal hdf pi-cent improvement in engine spaific 

furl wnsurnptxon. 

Stmpte cycle.-Bared on the minimal fuel economy 

improvement indicated and in consideration of the added 

hardware complexity and expected cost, the reheat cycle 

was eliminated in favor of the simple cycle, illustrated 

schematically in figure 8. The system, as illustrated in 

figure 8, is recovering exhaust heaf from the TC diesel of 

figure 1. The cycle performance shown. including final 

stack-gas temperature, results from an optimization of 

hearexchanger sizes in consideration of cost as well as 

performance. The total power output (diesel plus 

Rankine) considered against the did fuel rate indicates 

an sfc of 0.165 kg/kWh (0.271 lb/hp-hr); a 14-percent 

improvement ova that of the diesel core (see fu. 1). 

Figure 9 is a layout showing the diesel/steam system as 

it might kx installed in a truck. The major features an the 

condenser assembly replacing the now absent radiator in 

front of the adiabatic diesel, the heat-recovery steam 

generator replacing the truck muffler with similar backpressure and muffling characteristics, and the power 

module which includes the expander and water 

feedpump. The expander is a twocylinder piston device 

with design features similar to an earlier sixcylinder 

automotive engine (ref. 10). The expander operates a! 

diesel speed with a chain drive to the dial flywheel drive 

gear, thus avoiding the need for the multistage gear box 

assodated with reduction from turbine speeds. 

A piston expander was selected for the steam system to 

avoid the inherent low efficiencies associated with small 

axial-flow turbines. Water (steam) exhibits a high specific 

enthalpy; thus flow rates per unit power output are low. 

For low power levels the result typically is a very small 

turbine with the axial blades so short that normal tip 

clearance losses become relatively more significant, 

ruulting in poor overall efficiency. 

orgall& RanLfw System 

The use of organic fluids in substitution for water in 

Rankine power systems originated in low-temperature 

systems where the particular characteristics of certain 

organic fluids offer a performance advantage. Organic 

Rankine systems were the subject of developmental 

efforts as a potential automotive engine in the early 

1970's (ref. 9). The most recent automotive effort has 

Heat-rmry 

6n OC fl24U%: s$am qmeratw 

130P Qhr 12g86 Ibr'hrl ./ . Air inlet 

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RanMne 3~ ~i ( 9 

Total 275 W CH 0699) 

- flm rate 

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Water 

 pump 

Fum 8.-Digsei/m ~ystcm schematic. Rankine Fyde power output is bscd on Wway operaLiqn wilh condarser fan dccia. 
Skdm @%@fabr --fi 

Boiler fedpump -\, 

\ 

.----- 

\' ~aec transb care 

(Chaln drive) 

been the DOE-sponsored Truck Bottoming Cycle 

Program (ref. 11). 

The organic work fluid for the DOE Truck 

Bottoming Cycle Program was Fluorinal-85 with a 

maximum fluid working temperature of 288 "C (550 OF). 

The heat source was 482 'C (900 "F)exhaust gases from a 

conventional watercooled diesel. The program was 

completed in 1982 wi*a a highway ts series showing 12 

percent fuel economy improvement for the bottomingcycle equipped truck over an otherw~se identical truck. 

The advent of higher exhaust-gas temperatures 

assdated with the adiabatic diesel indicated a need for 

an organic working fluid with higher temperature 

capability. The fluid selected for this adiabatic 

diesel/organic Rankine conceptual dcsign evaluation is a 

mixture known from an early automotive Rankine fluids 

study (ref. 12) as RC-1. The specific conceptual design 

data and system performance information used in this 

comparative evaluation are based on a 1B3, DOEsponsored, RC-1 bottoming cycle study (ref. 13). 

The RC-1 effort of reference 13 included dyn3Jnic fluid 

loop testing to evaluate the thmal stability of RC-1 at 

temperature levels of interest for the adiabatic diesel 

application. Approximately 500 hr of testing were 

completed at each of three temperature levels; 371 'C 

(700 ", 427 "C (B00 'F), and 482 "C (900 OF). The test 

results indicated no evidence of thermal degradation of 

the fluid. 

High-performance cycle.-ln a manner somewhat 

analogous to the steam systems, the RC-1 power cycle 

can be configured either as a simple cycle or a highperformance cycle. In the high-performance cycle the 

liquid flow is dvided; part is vaporized in a diesel exhaust 

heat-recovery heat exchanger and then expanded in a 

high-temperature turbine while the remainder is directed 

to an enlarged regenerator for vaporization at a lower 

temperature by the supcrheated vapor leaving the hightemm;ure turbine. 

A preliminary analysis comparing the simple and the 

high-perfcrmance cycles indicated that over 85 percent of 

the potential fuel economy improvement is provided by 

the simple cycle configuration. A comparison of the two 

cycles indicated considerable additional complexity and 

potentially much greater cost associated with the highperformance cycle; that is, addition of the second vapor 

generator and turbine stage, increased control: to 

accommodate the flow split, and the fact that optimum 

performance involved a 66 "C (150 OF') higher operating 

temperature for the RC-1 working fluid. 

On the his of the relatively small performance 

difference and a reluctance to incorporate the more 

complex and costly system in a truck application, it was 

decided to emphasize the simple cycle configuration in 

the more detailed evaluation. 

Sfmple cycle.-Figure 10 is a schematic of thc simplecycle RC-1 system with fku conditions noted for 
Air htei lr; e;'? 2564 

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Pt9 bplhr W17 Iblhr), 

6.0 MPa 0370 pis); 

lssOCa7On 

JI L uJ2 -rI 

Dhl 236H 017 hp) @ Rankine Taw ZiPW 5 tW ( OTShpl I hpl 

Fm 10.-DicuVRC-1 organic systan -. Rankk cycl pnvu output is Won highway operation wiLb condenvr fan dcclutchcd. 

I 

I operation with the TC diesel of figure 1. The organic 

j (Rankine) system power output of 43 kW (58 hp) 

produces a redung sfc of 0.162 kg/kWh (0.266 lbhp-hr), 

which is a 15-perm improvement over the core diesel 

engine. 

The RC-1 fluid system with a significantly higher 

weight-flow rate and thus lower specific work in the 

i expansion process presents a more tractable single-stage 

! turbine design problem than the paranel steam system of 

I figure 8. The design used involves an 89 mm (3-1/2 in) 

diameter rotor operating at 55 000 rpm. The turbine 1 design thermal efficiency estimate of 77 permt is based 

1 on design experience with previous organic turbines (ref. I 14). The turbine gear box ovaall efficiency is 71 percent. 

Figure 11 illustrates the RC-1 system installed in a cabover diesel truck. In a manna similar to the steam system / Layout (fig. 9), the heat-rocovay heat exchanger (vapor 

j generator) replaces the truck muffler In a vertical position 

, 

4 

behind the diesel. A unique concept illustrated in figure 

11 is the rear-mounts condenser with air mp 

extending over the top of the twk cab in a manner 

similar to many diael engine air-intake system. The 

turbine gear box mata with the diesel at a standard lmD.3fU 

powa takeoff interface in the engine flywhed housing. 

The total speed reduction from turbine to diesel 

crankshaft is 29:l. FlguR 11.-WorSaaif ay$m Law. 

1 
The Brayton systems considered in thh report are all 

opencycle systems with air as the working fluid, and thus 

arc free from the high-pressure working fluid 

containment problems asodattd with the closedcycle 

steam or organic Rankine systems. In term of 

minimu any additional complexity, the air Brayton 

system could be considered as a next logical heat-recovery 

step beyond the turbocornpound diesel of figure 2. The 

Brayton cycle system is distinguished from the 

turbocompound system by the incorporation of a major 

heat exchanger as part of the cycle. The role of the heat 

exchanger is revd in the two major variations of the 

Brayton cycle as used for heat recovery. 

Sllb4tmosphalc BrPyton cycle.-Modification of the 

turbocompound system with a downstream heatrejection hea! exchanger (gas cooler) and compressor 

(exhauster) changes the system to a subatmospheric 

Brayton cycle configuration, as illustrated in figure 12. 

The w of a gas cooler and exhauster reduces the back 

pressure on the (turbocompound) power turbine to below 

atmospheric levels; thus the term "submospheric" 

configuration. The lower backpressure in the 

subatmospheric Brayton cycle effectiveiy increases the 

pressure ratio and power (heat) extraction of the power 

turbine as compared to the turbocompound system of 

figure 2. The key to cycle net power performance, 

however, is in the compression work required to exhaust 

the gsses to the atmosphere. Compression work, in turn, 

is keyed to the gas temperature at the start of 

compression and thus to the design of the gas cooler as a 

heat exchanga. 

The impact of heattxchanger design on the 

performana of the subatmospheric Rrayton system was 

evaluated parametrically (ref. 15) in comparison with a 

pressurized Brayton system. In the pressurized system the 

tole of the heat exchanger is the more traditional role of 

heat recovery; that is, heat is recovered from the exhaust 

gas and transferred to the power cycle. As illustrated in 

figure 13, pressures in this power cycle loop are typically 

above atmospheric; thus the term "pressurized" Brayton 

cycle. 

The Brayton zycle comparative anal: sis indicated that, 

in terms of net power recovery, the subatmospheric cycle 

is superior to the pressurixd cycle only if the heat 

exchanger (gas cooler) is sized such that the exhaust-gasto-ambient-air (approach) temperature difference is less 

than 45 'C (80 OF); or 64 'C (115 OF') if the cycle 

configuration includes one stage of compression 

intercooling. 

The dependence of the performance of the 

subatmospheric cycle on the very close approach 

temperatures in the heat exchanger creates two problems 

relative to the cycle hardware: (1) heat exchanger size and 

cost increases asymptotically at the closeapproach 

temperatures, and (2) the cold end of the heat exchanger 

is subject to high 'whg rates and potentially subject to 

acid corrosion. 
Information available from a previous experimtntal 

program Involving Mghcffwtiveness heat achangm 

operating in a truck diesel exhaust environment indicates 

that fouling of heatcxchanger surfaces is a significant 

problem (ref. 16). Fouling is generally considered to 

occur more rapidly at lower desel arhaust-gas 

tcmpemurcs. The accqxed practice has batn to avoid 

uroessive fouling rates and/or acid corrosion by 

rmhmhhg exhaustgas temperatures in the heat 

exchanger hve 149 'C (300 'F). 

Figure 12 illumes the Impact of clcse-approach 

temperature on the fluid temperatures in the 

subatmosph~c cycle. The 45 'C (60 'F) approachtanperm difference combined with the 29 'C (85 OF) 

ambient air temperature indicates a fd orhaustw 

tempcrature of 74 'C (165 '0, obviously well below the 

149 'C (300 OF) guideline mentioned above. Note that the 

same approach-tempaatun criteria applied to the 

pressurized cycle (fig. 13) docs not similarly imply low 

exhaust-gas temperatures. The nason for this difference 

is the cycle configuration in which the lowest temperature 

m that heat aFcfianSa is annprcaor didurge air ED 225 'C 

(437 'F). 

Ultimately, a decision was made to pick the pressuized 

Brayton system over the subatmospheric system. The 

deddon was based on the fact that any pos tive 

performance margin for the subatmospheric systera is 

keyed to a progressively larger and more expensive Ireat 

~as~ellasto~aylo~exhartst-gastemperp~~es 

Alr b, , 

ZPOCW 

that aggravate the fouling and add corrosion problem in 

the heat-rejdon heat exchmger. 

Press&& Bnrytorr-AGT ada,protlon.-The Brayton 

systems depicted in figures 12 and 13 are simplecycle 

mnflgura!.ions in that they represent the simplest 

operating form for the Brayton cycle. Lowcst capital or 

initial cost typically is asxkod with the simplecycle 

configuration. In an attempt to establish the low^ 

podble capital mi, a simplecycle, prcsurized Brayton 

system based on a minimum modification adaptatitn of 

the DOE/NASA AutomotiveGas-Turbine (AGT) engine 

was investigated. (Information provided by D. Evans and 

R. '?hnsen, NASA LRwis Research Center.) The 

mk~xum modification approach to adaptation involves 

some performance compromise, biu seeks to establish 

lowest capital cost by use of AOT components 

anticipated to be produced a! high rates and low cost for 

future automotive applications. 

The AGT adaptation to diesel exhaust-heat recovery 

mtm on removal of the AGT's fud combustor and 

modificarion of ducting as required to permit the rotary 

regenerator to function as a heat-recovery heat 

exchanger. Use of the automotive rotating regenerator as 

a heat-recovery heat exchanger imposes performance 

penalties on the Brayton system due to leakage past the 

rotating-to-fixed element seals and the relatively Lgh 

flow resistance or pressure loss &atd aith the 

regenerator core. Performance parameters for the AGT 

adaptation are illustrated in figure 13. Yhe 19 kW (t6 hp) 

Turbochargar 671% w%; m-ry 

heal eahm ger 

'.. i..* ..-- 

W2 % 0158 OF), POC 

QdMR (MmW 

- mrrlur - , d. ..p a .- --. -,-- Its), 
net output of the Brayton system rsuh in a compound 

engine system sfc of 0.177 kg/kWh (0.291 Ibhp-hr); a 

modst 8-pacent improvcmm over the aue did. 

PnrrrPfred Buyrod frrta#rdcd q&.-In addifion to 

the simple-cycle or AGT adapulion approach desaibed 

prrvioudy, the pressurized Brayton system was also 

investigated (ref. 15) on an optimited "dean-sheet" basis 

(i.e., without the restriction of any pmious hardware 

desigu) using an intercooled cyde in an attempt to 

improve pafonnance. The cyde is itlustrared in figure 14. 

Incorporating one stage of interrwluq into the 

Brayton cycle nut only reduces the average air 

tanpaatwe dining compression, it also allows for 

wtly better compressor efficiency because the 

pressure-ratio requirement is split berween the two 

-heels. Tbe negative aspect 2 f the intacrzoled cycle is the 

added hardware; that is, a second comprmr wheel and 

the air-to-air intercooler or heat exchanger. 

Ihe optimved dean-sheet design approach for the 

intwed cycle included use of a fixed recuperator as 

the heat-recovery heat exchanger. Analysis of this cycle 

indicates h, the net power outprrt is 27 kW (36 hp) and 

the compound engine sfc is 0.172 kgflrWh (0283 Ibhphr); 

a lOpacent improvunent over the co, : diesel. Figurc 15 

iIlmtmtes a typical hardware packagmg amangement for 

the intamled cycle. 

Power Cycle Comparison Factors 

A comparative evaluation, as indimed in the trtle of 

this report, requires a comparable basis from which the 

evaluation is conducted. In this regard, the initial don 

of this report established the adiabatic diesel baseline 

which defined the diesel exhaust conditions which are the 

heat source for the heat-reavay power cycles. With a 

consistent adiabatic did heat source defined, the 

alternative powa cycle configurations were developed as 

illustrated in the second section of the report and the 

performance ctiaracteristics of each were &.mated. lEe 

configuration effort included component descriptions 

adequate for cone cost estimates which will tx 

introdcod in this section of the report as various 

performance and axa comparisons are made between the 

various alternative-powercycle configurations. 

Spedfic fud consumption (sfc) is the fundamental 

engine perfonnance paratneter that influences truck fuel 

economy. In the case of a diesel with heat-recovery power 

cyde, the performance parameter of inter&. is the sfc of 

the compound engine. Compound engine sfc is the did 

fuel me divided by the combed pwer output of the 

diesel reciprocator plus the heat-recovery pwer cyde. 

Fv 14.-Dkd/prasunmJ (mtrrcoded) Breytan system srhanatic. badd byton ornpLd travd K @MY 0perati~ 

imcroodcr fan an boi speed. 
F&-reCoF.ery 

heal exhangar \:J-~ tnM Dd 

Compound engine sfc was initially waluated in a 

parametric phase in which all four of the did 

configurations, TC, TC/A, TCPA, and TCPD/A, were 

evaluated separately with each of the alternative powa 

cydes. The paramefric analysis indicated tha! from an 

overall sfc viewpoint, nonaftercooled diesels are 

preferred as the core for a compund engine system. As a 

did only, the nonaftercwled engines show a shghtly 

higher sfc accompanied by a relatively higher exhaust-gas 

tenpaazure (fw. 1 and 2). In the ampound engine 

configurations, induding efficient exhaust-gas ha! 

recovery, this sfc trend is reversed and the conmons 

with the nonaftemled core show equal or better sfc 

results. Acmrdmjjy, the emphasis in the comparison 

efforts will be on the use of the nonaftercooled diesel 

engines CTC and TCPD) as the core units. 

Figure 16 is a summary of the sfc resuhs illustrating the 

diesel baseline data and the performance of the TC and 

TCPD diesel-plus-heat-remvery power cycle compound 

engine systems. The results illustrate an appromely 

TCI TCWl TCI TCPOl TCI TCPOl TCI TCPDl 

ACT ACT Bw- Bray- Ueam Steam RC-1 RC-1 

brr bn Or- Organic gank 

Figure Id-!S@fk fucl owwmphn (sfc) for various dicwYaltanativewaw combinatiw* 
! The purchase price or initial capital cos of a systern is j a major fanor in the determination of its relative 

desirability. In the case of a fuel-zffiaent engine system 

the buyer is expecting annual fuel expense savings that 

will pay back the ~ncremental or eara capital cost within 

some fraction of the useful life. The balance of the 1 ownership mod thus represents a net gain from the 

. - purchase decision. 

i Unfortunately, the cost estimates for advanced systems 

-a i are typically less exact, or at least less uniform, than the 

performance estimates. Two of the major uncertainties 

involved are the anticipated production rate as it impacts 

unit manufacturing cost and the pricing environment 

which affm the markup ratio to sdhg price. i; ihdudw rare.-Figure 18 illustrates the relative 

impact or sensitivity of unit manufacturing cost to annual 

production rate. The points shown reprat mst 

estimates of a Rankine heat-recovery system developed 

for three production rates (ref. 13). The 10 000-unit rate . . !: I selected for use in this audy is based on an assumption of 

I! 10 percent penetration into a class-8 diesel truck market 

- 

* 

C 

I I 

10 1m 

Annual prductim rate, unb 

19-percent range of sfc values between the worst baseline 

diesel UC) and the best disei/alternative-powerqcle 

configuration (TCPD/organic). The appeal of the 

turbcampound concept is wident in that it is dearly the 

simplest form of heat recovery but still produces au 

attractive 6 percent sfc gain (TCPDIA diesel as opposed 

averaging 100 000 units per year. Class-8 diesel sales over 

a 6-yr period from 1977 to 1982 have ranged from a high 

of 158 000 units to a low of 67 000 units (ref. 17). Class 8 

coven trucks from 15 000 kg (33 000 Ib) gross vehicle 

weight up to the legal limit of 36 000 kg (80 000 Ib) and 

trucks used in all types of senice from local to long-haul. 

It is considered that heat-recovery cycles would be 

attraaive to the fraaion of the class-8 market 

represented by the h& gross weight trucks in long-haul 

SeTYj~. 

Mmkup.--The markup ratio from unit manufacturing 

cost to selling price varies depending on market 

conditions and on the bargaining strength of the buyer. 

For this study a ratio of 2-to-1 was assumed as 

reprsentative of average market conditions and the 

bargaining strength of a fleet purchaser. The resulting 

fleet price is considered to be consistent with the did 

engine price levels established in a previous section of this 

report. 

to TC/A diesel). The diesel-plus-altema,tive-powercycle 

results, however, indicate a potential for up to 12 perm 

more improvement ova the TCPD/A ~3formanoe. 

figure i7 shows how the sfc results of fuure 16 

combine with the fuel use cordation of figure 5 to 

produce an annual fuel use reduction for each compound 

ACT ACT Bray- Bray- Steam Stsarn RC-1 RC-1 

bn bn Or- Organic Fk 

Fv 17.-lncrrmcntal annual fuel SB~ mutt@ from usc of varbu WMve power cycla based on tbc TCPD/A diesel a the 

basdim wf-n Bnd 161 aT) km (100 CCQ mi) annual w. 

14 
were used witb appropriate adiabatic did prices to 

develop an overall price for the various ampound engine 

syst-. 

c0mptn.d ardfrrc pkr.-Compound engine prices 

were developed by combining the appropriate did core 

prices from figure 6 with the altanative-powercyde 

system prim from table I1 and then scaling as required to 

refled the compound engine price at a dected common 

power level of 261 kW (350 hp). An example of this 

overall prdure applied ;o the TC/steam ampound 

engine system is as follows: 

. 

i 

\ - 

, 

TC died $14 000 

Steam cyde + $6070 

Compound engine $20 070 

C'prZw e&tares.-Table Il is a tabulation of the 

estimated capital cost or purchase price of each of the 

several alternative-power-cycle systems under 

consideration. The data shown were developed 

independtntly in each case by an advocate for that 

system. In general, all prices were developed by 

The resulting compound engine system, as illustrated in 

figure 8 ha a combined power output of 275 kW (369 hp). 

At this point, the 0.7 errponen! pri- factor 

described earlier is applied to adjust to the common 

power level. The resulting price is then compared to the 

price of an equal power level TCPD/A did taken from 

figure 6: 

Note that the incremental difference shown for the equal 

power level compound engine system i . onsiderably lers 

than the $4070 price difference between the 

turbocornpound and steam system indicated in table 11. 

The comparison at equal power levels acknowledges the 

higher efficiency and thus greater power contribution of 

the steam system. 

eshmbg component unit manufacturing costs and then 

applying the mmrnon 2-to-1 markup factor to reflcO <I uUalIedp-toafleetpwchaser. 

1 TABLE 11.-PRICE ESnMATES FOR ALTERNATE POWER CYCLES 

b, The data in table I1 were hedoped on the bads of a 

production rate of 10 000 units per year, except for the 

AGT price which is based on a rate of 300 000 units. As 

discussed previously. the AGT system is assumed to 

represent a minimum modification of .e automotive 

@arsenger car) gas-turbine engine. hrdmgly, the 

AGT price gtimate reflects significant benefit from the 

automotive production. Figure 18 indicates that unit 

costs at the U)0 000ynit production rate typical of , - 

- 

.. . - 

.rr; 

E- 

,, - 

automobile produdion lines may be only about onethird 

the axt of a comparable unit produced at the 10 000unir 

rate. 

Table I1 inclubs notations on seleaed physical 

parameten that are considered to be useful in 

rationdhq some of the more significant component 

price differences heen systems. Thc heat exchangers, 

in addition to bang sized by the surface area noted on the 

table, can be charaaerited as either plate-fin or finnedtube type. The plate-fm type, which for this study is 

unique to the intercooled Brayton system, is typically 

more compact and correspondingly les expensive than 

tbe finned-tube type. 

The lxmmn line on table I1 is the indicated total price 

for each system. WW the Limits of aauracy arpected 

for a conceptual design study, the prices shown are 

considered to be reasonable. Accordmgly, these prim 
Fif.urt 19 illusrates the incremental difference 

Ci-) in capital cost arsociated with each of the 

candidate ahanative power cyds as applied to the TC 

diesel and also as applied to the TCPD died. Note that 

the lower exhaust-gas temperature of the TCPD did 

(fig. 2) results in downsizing of the ahanative power 

cycle Price eaimats for these downsid units were 

dd from the data of table I1 using the 0.7 aponent 

tecbrique dgcribed pmiously. The AGT sygan was an 

en~r,$ion for whid~ cost scabg was not considered 

~.iopriate. 

Afabrt- and rrprcfr cost mm@son.-The earlier 

disc-&on of the basehe adiabatic did maintenance 

and repair cw~ served to illustrate the importance of 

these msts in high kilometem (miles)-per-year long-haul 

truching applications. The various synem schanatic 

diq; ams preser~ted in the cyde configurations seaion of 

tkG ,eport show that in aich case the alternative power 

~ydr amolmts to a rather significant sulxpstern addition 

to the did powerplant. The implication is that the 

mhaenanoe and repair burden of these subsystems will 

also be significant. 

The capital price breakdown in table I1 indicates that 

the hardware of the alternative power cycles can be 

segqaed as to prime mom, heat exchangers, and 

mntrols/fluids. A similar approach is taken for the 

maintenance and repair a&s as illwtmed in table 111. 

The oosts shown in table I11 were developed by NASA 

(,exceptions win be noted) with emphasis on the relative 

costs from system to sysrem. In a manner nmilar to the 

diabatic did maintenance and repair estimates, the 

1 om cycle estimate are bad on a levelized annual 

payment for a lifa' & &maintenance and repair cantract. 

A srarting point for the maintenance and repair cost 

estimates was a figurr of 53% received from Cummins 

hghe Company pasonnd as a preliminary estimated 

annual expense for tke turbcmmpound power turbine 

and gear train, including fluid coupling. From this basis, 

the prime mover package estimates for the other 

alternative power cycles were developed by incrementally 

adding costs in rehionship to perceived complexity. La 

the case of the steam and organic systems the need for a 

variable stroke fluid feedpump war also considered in the 

prime mover mst category. 

ffil A~T Bray- Bra)- Steam Sb RC-1 RC-1 

tan bl Or- OrFm 19.-Inaund pna of various dieul/atumativepowucyde 

combinarions normalltcd to 261 kW (350 hp) total power. The 

-priccisWmthTCPD/AdicsdpriceofS16U. 

TABLE LU.--MAINTENANCE AND REPAIR COST ESITMATES 

FOR ALTERh'ATE POWR CYCLE5 
Heat exchanger maintenance and repair casts were 

bad on the radiator and fan data that had earlier been 

subtrac!cd from the state-of-the-art disel contraa repair 

program (ref. 8) in the pmuss of developing tbe 

adiabatic did mainterma and repair ertimates of 

figure 7. The WJ radiator and fan estimate was adjusted 

downward for the simpler gas-to-air beat srcfLangas used 

in the Brayton system. The AGT mzary regenerator with 

asodated regenerator sca.Ls was a spcclal case asig~ed 

the full E41X) w . This is consided to be consistent with 

the $340 cost estumled in the NASA study of the AGT 

heat-nxovay system. (Information provided by D. 

Evans and R. Johnstn, NASA Lewis Researcfi hta.) 

The maintenawe and repair cost basehe for the 

controls and fluid inventory megory was adapted from 

the estimate developed in reference 10 for the controls of 

the steam system. This cost was increased by SlOO for the 

organic system to allow for replacement of the organic 

fluid charge as reqnred. 

The maintenance and repair axt gtimates from table 

IIl were combined with the adiabatic digel maintenance 

and repair cost estimates of figure 7 in developing 

mairrtenance and repair msts for the various compound 

engine syjtems. An example of the prm for the TCplussteam system is as follow: 

TC diesel $3335 

Steam cycle + $1500 

Compound engine $4835 

Note ihat these ~osts are for a compound engine rated at 

275 kW (369 hp) as illustrated in figure 8. Consistent with 

the cost estimating lines in figure 7, an qnent of 1.14 

was used in adjusting thtse costs downward to refled a 

common engine size of 261 kW (350 hp). At this point, 

the esthmed maintenance and repair casts for this 

compound engine are compared to similar costs for the 

TCPD/A engine: 

TC plus stcam $4552 

TCPD/A - $4015 

Increment S 537 

Note that the $537 increment differs from the result 

($1 50 - $350 = $1 150) obtained by working directly with 

the data in table 111. The reason here, as with the capital 

cost data, is that the comparison at equal power levels 

acknowledges the higher efficiency and thus greater 

power contribution of the steam system. Figure 20 

illustrates the incremental annual maintenance and repair 

costs assodated with application of the various 

alternative power cycles to the TC diesel and as applied to 

the TCPD diesel. The figure shows the lwW annual 

cost for a 7-year maintenance and repair c~nuaa and 

161 00 km (100 000 mi) annual use. 

AGl rcr Bny &my- Swa Stxu RC-1 Pf-1 

bn bl Or- Or- 

*@ 

Fw 20.-lnrrmnnal annual maintenana End repair axl3 for 

vrniom dKsYakanarive-poon& armbinations "d to 

261 kWOWhp)tatalpo~.~~priocts~cmtbc 

TCPD/A dicPi cast of W)15 pa pu. 

l%e prospects for widespread implementation of any 

of the diesel/altanative-powercycIe systems are entirely 

dependent on the ability of the systems to compete on an 

economic basis. One expression of the economic merit of 

a system is the operating time requmd to pay back or 

recover [he initial investment increment. Recent 

marketing studies of the Organic Rankine Cycle System 

(ORCS) for long-haul truck applications have indicated 

that 2.5 to 3.0 yr is perceived as an acceptable paybask 

time (ref. 18). A separate survey of heavyduty trucking 

engine requirements (ref. 19) concluded that 1.5 to 2.0 yr 

is a desired payback time. 

The conoept of payback time as an economic measure 

is predicated on a simple reward/risk relationship which 

is apparent from the dewon of the payback 

rekionships EU follows: 

risk capital Payback time = annual savings 

Net reward - (lifetime X annual savings)-iisk capital - risk risk capital (2) 

E. 

lifetime 

pa-k time - ' 

Figure 21 illma the rdationship of payback time to 

nu reward/risk ratio for an wed 7-year life typical ~f 

truck diesel engins. The figure indicates that 3 yr as a 

maximum aooeptablc payback period equates to a 

Lifetime net reward/risk ratio of 1.33; that is, payback of 

the inttial invff~ent plus 51.33 nu reward on & $1.00 

mvsted. Payhack in 2.3 yr (28 mom) indiauo) a act 
reward ratio of 2.0 or a $2.00 net reward for every $1.00 

risked. 

For payhack calculations on the various diesel-plusalternative-powercycle combinations in this study, the 

risk capital is the price premium or incremental price 

increase over the competing TCPD/A diesel. The 

incremental price of the E/steam system over a 

TCPD/A diesel is illudrated in figure 19 as $2503. This 

dollar number is then modified by a factor of 0.85 to 

reflect an assumed 15-percent salvage value in the 

equipment st the end of the 7-year lifetime. The resulting 

risk capital amount is Q128. 

The annual savings is the fuel (qxme) redudion 

minus the increased maintenance and repair burden 

associated with a more complex total engine system. All 

the values are incremental changes with reference to the 

TCPD/A diesel. The annual savings afforded by the 

TC/steam system can be calculated as follows using the 

fuel price of W.32fliter ($1. Wgal): 

Annual fuel eqmw reduction (fu. 17) SS92 

Annual increase in rnainmamx and 

repair - (f'. 20) 

AMU~ savings (net) 

The payback time for the TC/steam system can be 

calculated (eq. 1) as Q128/($455/yr) =4.7 yr. 

Unfortunatdy, this r& Indicates the system is not 

economically attradive at currat fuel prices. 

Figures 22 and 23 illustrate the paylmck results for the 

various candidate ah~power-cyclc systems as 

Fv 22-Paw tiw for various TC &d/ahanaffw-pmucyde am^ the TCPD/A dkd ar a basdine for paybask 

-pan'son. 

~PPW to the TC did and as applied the TeD 23.-Psmk hrn &+sd/*cr~n-po.*rr- diesel. The payback results are presented as a function of cyde axnbtnadons with he TCPDIA dkvi as a bardioe for payback 

fuel price, asamhg an arbitrary increase in fuel pria not -. 
accompanied by an increase in cquipmm and/or 

maintenance costs. Ln each case, the basis for payback 

comparison is a TCPD/A died of equal pow. 

The payback results indicate that, at the current fuel 

price, none of the candidate altem&tive-powercycle 

systems is economically attraQive as a replacement for 

the TCPD/A died. As fud price increases, the fust 

systems to become attracrive are the Rankine cycle 

systems followed by the intercooled Brayton system. In 

all cases, use of the TCPD diesel is more attractive 

(provides quicker payback) than use of the TC diesel. 

In an attempt to gain additional insight into the relative 

economic a@ of various engine configurations, the 

methods previously described were used to calculate the 

paybark of a TCPDIA diesel versus the TC/A diefel. 

The results, based on the adiabatic diesel data in this 

repon, indicate that the more fuel efficient TCPD/A 

died (versus TC/A) crosses the 3-year payback 

threshold at a fuel price of S0.23flirer 0.87/gal) and at 

current fuel prices would payback in an attraQive 1.92 yr. 

Cowe of Stepm Versus RC-1 Orgnntc Rtdd far * Cy* 

The Me power cycle is identified by the fact that 

the state of the working fluid changes from liquid to 

vapor and then back to liquid at specific locations in the 

dosed-loop system. Because a change of state is involved, 

the unique characteristics of the working fluid in tams of 

the liquid-vapor phase interface (vapor dome) -me 

important. 

Steam is the seleded workmg fluid for the vast 

majority of operating Rankine cycle power systems. Its 

characteristics of low cat and wide availability have 

made it the universal choice for direct-fired applications. 

For waste-heat applications, however, the various 

organic fluids have dved considerdon for their 

unique c~aistics that provide berter efficiency when 

coupled to a low-temperature heat source. The key factor 

involved here is the heat recovery eficiency which is a 

measure of the efficiency with which heat is exrrscted 

from the waste-heat sueam. 

Figures 24 and 25 are temperature as a function of 

enthalpy (t versus H) plots illustrating the steam Rankine 

and the RC-I organic Rankine configurations introduced 

earlier in this report. Included on each plot are the vapor 

dome charasteristics of the rapative fluid as well as line 

representing the dieselexhaust weam of the TC 

adialratic diesel as used in this study and the bwertwpaature exhaust of a anventionid watercooled 

diesel engb-e. 

The Phapc of the vapor dome influences heat-recovery 

eff~ency via the Idon of the socalled pinch point, or 

point of minimum temperature diffaence between the 

exhaust-gas stream and the cycle working fluid. W~th 

RC-1 organic fluid, the vapor dome is skewed such that 

'-RC-1 Wlrq 

fluid In eycle 

Vteatadditton) 

Exhaust pas fmm 

mr ---\ -.,/ 

Fw 25.-Heat addih~ profile of Rankint cycle. 

the pinch point is at the low temperature end of the heat 

exchanger (fu. 24). This location allows for efficient 

cooling of gas streams regardless of their starting 

temperature. In the steam system (fig. 25) the pinch point 

is in effed a pivot for the exhaust-gas cooling lines of 

various tempemure origins. The result is that the hightemperature exhaust for the adiabatic died is actually 

cwled (heat mvered) to a lower frnal temperature than 

the exhaust srresm from the conventional water-led 

diesel. This meam that the steam system's heat-recovery 

efficiency improves sigmflcantly when used with the 

adiabatic did. 

Another item illwed in figures 24 and 25 is the fact 

that the steam system has been designed for a 

sigmfjlcantly wida pinch-kmpaatute difference than the 
RC-1 organic system. The phch-tempaature difference 

is a major factor impacting heat archanger site. The 

wider pinch temperature us& in the steam system 

rep- a compromise in heat-addition effidency and 

also acaunts for most of the price difference between the 

steam and RC-1 organic systems as shown in table U. 

Reducing the pinh-temperatun difference of the 

steam system to equal that of the RC-1 organic system 

would allow the don ofheat from the exhaust-gas 

sueam down to a temperature of approximately 163 'C 

(325 'F). The additional heat recoveted would boost the 

power level of the steam system to essentially equal that 

of the RC-1 organic system, but the corresponding 

requirement for increased vapor generator surface area 

(table 11) would also boost the price to near that of the 

RC-1 organic system. Thus, the steam and RC-1 organic 

Rankine systems can be my equal in perfonnance 

and price if both are configured with the same dm 

philosophy for application with an arfiabatic diesel. 

Cydc~to~FL.harrPrTemperPtme 

As discussed earlier, the adiabatic diesel is an advanced 

mnoept just entering the early developmental stages. 

Amrdmgly, the engine amfiguration and operating 

parameim are not fured. Among the parameters subject 

to change is the exhaust* temperature. For this reason 

it is mie to establish the sensitivity of the various 

alternative power cydes to variations in temperature of 

the did exhaust gases. 

The most immediate and dim impact of exhaust-gas 

ternpaatme on the altanative power cycles is the 

variation of power ouiput. Typically, the alternative 

power cydes are capable of cooling exhaust gases to a 

final stack temperature in the range of 1% to 200 'C (300 

to 400 'F). Based on 150 'C (300 OF) as the fd stack 

tempemure, increasing the exhaust gas starting 

mpaature from 538 'C (1000 'F) up to 871 'C (1600 'F') 

results in an 86-percent inaease in available temperature 

difference and thus, available energy. 

The extent to which the ewrgy available in highertemperature arhaw - an be lailized depends on the 

characleristics of the individual power cydes. Fwe 26 

illurrtratts the povm orrtput sensitivity of the RC-1 

organic Rankine system, the intercooled Brayton system, 

and the steam Rankinc system over a range of diesel 

exhaul-gas marting temperatures. Both the Erayton and 

steam systems display grater response to exhaust-gas 

tanpaature increases than the RC-1 organic system. The 

RC-1 organic systm is the least responsive system 

becaw of Mons on the maximum safeoperating 

temperature of the RC-1 organic working fluid. 

Prdiminary data from laboratory testing (ref. 13) 

indicates the RC-1 organic fluid may be stable at 

tanpacaurs up to 480 'C (900 'F). For pllrposs of this 

Fi Zb-Sendtivily of aLcmative power cycle pafurmance to did 

ahaw mpcrature. (Died Uhaua fbot ratc 1309 Lg/hr (28% 

bh).) 

study, however, an RC-1 organic fluid temperature limit 

of 400 "C (79 OF) was assumed to represent a safe 

margin against fluid deterioration in long-tam savice 

under field operrtting conditions. 

The steam Rankine and the Brayton performance 

projections shown in figure 26 assume that the cycle 

working fluid (steam or air) peak temperature will be 

allowed to increase as the exhaust-gas or heatsource 

temperature increase. Amrdmgly, the performance 

benefits from improvements in cycle efficiency as well as 

from the increase in energy available in the diesel 

exhaust. The steam cycle receives an additional benefit 

from improvement in the gas-tosteam heat-recovery 

efficiency associated with the higher initial temperature 

of the heat source. 

A simant point illustrated in figure 26 is that, in 

tenns of cycle power performance, the Rankine systems 

remain superior to the Brayton system within the 

temperature range examined. Regarding the choice 

between steam or the RC-1 organic as the Rankine cycle 

working fluid, the higher gas-source temperatures favor 

the steam system. Note also that the data in figure 26 

were developed from the specific conceptual design 

results for each cycle concept. As discussed in the 

previous section, the heat-recovery heat exchanger for the 

steam system is a ampromise for cosr at about half the 

surf- area of the corresponding unit in the RC-1 

organic system. The point is that the power performance 

curves shown in figure 26 are relative. Power 

perfonnance of the steam and RC-1 organic Rankine 

systtms would be approximately equal at the 671 'C 
(1240 'P) exhaust-gas tanperatwe level if the hea!- 

recovay beat admgus were of comparable size. 

For purposa of this report, the tam "ttechnological 

barrier" refas to major tahnological problems that 

would dun an appropriate industrial source from 

p& an otherwise amxczive dieul/heat-recovay 

power cyde systan. This definition excludes the more 

routine product development problems as beh~ the 

normal responsibility of the indllstry in bringing a new 

product to the market. 

In general, the environment and operating 

requirements of a heat-recovery power cycle are 

consisten! with state-of-the-art technology. Working as it 

does, with the heat rejected from a prime (diesel) cycle, 

the tanpaarure levels of the heat-recovery powa cycle 

are modest in comparison to the prime cycle. There are, 

however, tedmolugical barriers assoda2d with efficient 

recovery of the heat from the diesel exhaust stcam and 

tvhh the operation of the steam and organic cycles. 

Hat erdmg# fouling.-Effedive heat recovery 

within envelope limits of a typical truck configuration 

requires a compact heat exchanger with closely spaced 

pisqes. Previous experimmtal dts (ref. 11) indicate 

that such heat exchangers are quickly fouled by the 

paruculate laden did exhaust gaszs. 

There are some prdiminary data (ref. 1) on the 

adiabatic did that indicates particulate levels for such 

an engine may be dramatically lower than for 

conventional watercoded engines. An dcipated trend 

toward the use of heavier grade fuds, however, would 

Iikdy rrinvodua the fouling problem. Accordingly, gasside foultng of the heat-recovery heat archanger remains 

a prfmary technological barrier that influences all power 

cycles. 

Expaimental work has been accomplished (refs. 20 to 

23, and work is cominuing (ref. 23), in regard to the 

fouling problem. The srperimcmal investigations have 

related fouling rates to gag temperatures and metal 

temperatures in the heat exchanga. M&& of 

prevention and/or removal have included mot blowing 

and periodic water wash as well as a dfdeaning 

technique in which the cycle working fluid flow is 

interrupted for a period of time to allow md 

temperatures to rise to the point where soot dries and 

flakes or can be blown off. The latta technique generally 

u not amsidered to be applicable to orgdc fluid systems 

due to a risk of overtemperature daxmposition of any 

organic fluid residue remaining Ln the heat exchanger 

during the bigh-tempcrarure opaation. 

Ckgdc fbdd cAProct--The organic working 

fluid designated RC-1 is a dxtm of 60 mole pcrccnt 

pcntafluorobenzene and 40 mole percent 

hexafluorobmzene. The fluid presently has an 1 

cxpaimentaf status, having been examinad in laboratory 

tests but never used in a Rankhe powa system of the 

type desnikd in this report. 

Acocptance of the RC-1 fluid for on-highway power IS 

cyde applications would @ extensive experimental 

verification of the fluid charactaistics in tenm of 1 " 

thermal Wty under operating conditions as well as 

potential environmental and/or safety hazards. Only 

preliminary work has been performed in this area to dare i 

(refs. 12 and 13). ,' I 

&ewn system ftatLs protection.-The water working 

fluid used in the steam system cannot be proteded by an 

antifreeze, as is the water in a conventional engine 2 

1 

cooling system, because of the significantly higher 

temperatures ~nvolved in the steam powa system. 

Accordingly, the steam system could be subject to 

hardware failure kaw of wer fnezine during a 

shutdown paid. A 

'Re old-time steam cars avoided the freezing problem 

by a combination of limiting winter exposure and 

.I 

operrttion on a opencycle basis such that the system 

'.~

could be boiled dry prior to any period of cold weather 

shutdown. This approach is not acceptable for modan, i 

, , 

high-performance applications, however. -. , 

Relirmnary work indicates that propa sizing and i design of the bailer and condenser water passages can 4 

allow for safe drainage to a proteded sump during -1.. 4 

- shutdown. Confion of this design approach by an - 

, . 

appropriate system test is needd to remove this perceived 4 .: ' 

barrier. 

Sysra cw and wmplo;rlty.-The significant fuel A; 1 

economy improvement potential of the diesel/Rankine ,. 

compound engine srjtem is und~~utd. Widespread ,l 

implementation of the concept, however, is blocked by , -=:\

. - 

concerns for adequate economic and operational : 

payback in view of the very significant hardware cost and -4%. , , *st I 

complexity. 

A potential for meaningful reduction in cost and i 

I 

complexity of the adiabatic diesel/Rankine system may i 

lie in an hegrated engine approach. The integrated I 

approach involves the adaptation of one or more - 1 

cylinders in the did block to Rankine power. The 

integration effort would include design of the Rankine ! 

vapor generator for close cuuplmg to the engine within 

the engine compartment. Such an approach would 1 

require the services of both engine and vehicle designers. 

The comparative evaluation of alternative powa cydcs 

for adabatic did waste-heat mvery involved a 

compla integration of techdcal and economic factors. 

The dor technical basis for the study is a saics of heat- 
mvq power cycle conceptual including 

estimates of cost and design poh performance. Highway 

furl mnomy was then assumed to scale on the basis of 

relative changes in engine d&n point performima. 

The coatidence led of the comparison could be 

improved by increased emphasis on the economic and 

mission fadon. Spedc steps for improvement would 

indude miew of caphal and maintenance oosts by an 

indqmdent indusuy source, devdopment of engine-part 

power maps, and use of a valid truck mission-simulation 

md to evaluate highway fuel economy. 

With these limitations in mind, the study results 

indicate the followiag conclusions: 

(1) Fuel mver trucks incorporating significant 

nonengine fucl saving features duce the mission enagy 

Id and thus the economic justification for a heatrecovery power cyde. 

(2) An increase in fuel price improves the economic 

justification for heat-rewvery power cycles, but only to 

the men! that the fuel price change is not aanmpanied, 

or dosdy followed by, asrociated engine price inflation. 

(3) The economic justifladion for turbocompounding 

is strong. The TCPD/A engine, compared against the 

TC/A engine, showed a very attractive payback time a! 

current fuel prim. 

(4) The econoric justification for displacing the 

TCPD/A engine with a dieselheat-recovery power cyde 

engine is weak. An arbitrary fuel price increase, not 

accompanied by hardware cost inflation, is needed to 

provide economic payback within even the maximum 

acoeptable time period (3 yr). 

(5) The fuel economy performance and the economic 

judfiauion is kst where the heat-recovery power cycle is 

used (added) in series with turbocompounding; that is, 

rather than as a substEtute for turtiamrnpounding. 

(6) Among the heat-recovery power cycles studied, the 

Rankine cycle provides the best fuel economy; up to a 

12-percent improvement over the TCPD/A engine. The 

Rankine cycle also shows the best relative economic 

justification of the ha-recovery power cycles studied. 

(7) At the tmptrraure levels indicated for the 

adiabatic diesel, the Rank.int cycle working fluid can be 

either RC-I organic or steam rsulting in essdaily the 

same power perfonnance and economic payback. 

Regardless of which fluid is us&, the dsign criteria 

should k a compromise ktwetn performance and cost. 

(8) The diesd/Brayton systems are inferior to the 

diescl/Rmkine systems in terms of fuel aanomy and 

axmomic justiflcatian. The low eost diesel/AGT 

adaptation shows only minimal fuel economy 

improvement ova the TCPD/A engine. The opdmfLtd 

dlcscUBrayton system costs as much as a dieseVRankine 

system, but provides only blf the fuel eCOLlDmy 

improvantot. 

(9) The major technologtcal barriers associated with 

the heat-recovery power cycles are the indicated 

complexity and rsulting capital and maintenance ma of 

the systems. These factors seriously weaken the economic 

attractfveness of the systems. An additional barrier is that 

of diesel exhaust-gas foulin8 the heat-recovery heat 

erchanPa. 

Natiod Aerowcs and Space Adxumm . , 'on 

Lewis Research Cauer 

April 18, 1985 

References 

1. Brydt, Weha; d ko, Ray: TACO.WCummifin AdiabgtL 

E&uc Program. SAE Papr 830314, Fcb. 1983. 

2 Amalcrm Trucking T& lmlP%O. Tn&q . . Ascamam, Inc., Wahingm, D.C. 1PS1, p. 66. 

3. Brands, Michod C.; Wana, John R.; and Hochnc, John L.: 

Vchidc rating of Cumdm Turbocompound DIesd Engine. 

(CTRU74HCKXL Cummins Engm Co.; NASA CoanaQ EM-78- 

CM-4?36.) NASA-CR-1 JW, 1980, p. I I. 

4. Crmr, Rich: Rkte Reas: 'he Dominant RnL m US. Tntdiag. 

Cmmd Ca.na Jwmd, w1. 140, no. 6, Jlmc 1983, pp. 115, 

116. 120-123, 126129. 

5. Gdoo Remiss 10 MPG 1990. Cbmmdd ("Min Journal, 

vol. 140, no. 6, June 1983, p. 54. 

6. Cummins Ttn Rig Negn 10 MPO. Heavy Duty Trusking, val. 63, 

00. 1, Jm. 1984, p. 22.. 

7. Farsfblldy Ttn of CompormQng the Imernal Comb4tton Enefne 

for Automotive Vthida, TaJk 11. COa2690.1, TbamtElccnon 

Curpomption, Wahham, MA, 1975, pp. 10-15. 

a. Conrrrnst RcpairqCrE& for Rofil. EUllain 3382459, Crmrmins 

En&c Co., Colmnbus, IN, Jan. 1980. 

9. Ludxer, Stcphcn; 8nd Rmaa, Roy A: An Arrarmmt of the 

~&loay w ~nstna for wm. ERDA n.M, 

Apr. 19n. 

10. Poulfn, E, a al.: Steam Bonomlng Cycle for Adlabgttc Mrsd 

we. (FMI-NAS8m, Fusta-.Wla Asdam; NASA 

GxxW& DEN3-3W.) NASA CR-168255, 1B4. 

11. DiBdla Fmnck A.; Imamo, Lum R.; wd Kopkm, .Hlchad D.: 

Laborrtlory d On-Highway Tc- of D&si Omnk Rankine 

Compormd big-Haul vejnlcle Engine. SAE PPpCI 830122, kb. 

1983. 

12. MiIln, D.R., d al.: Opimum Working Fld& for Arnodve 

Ranldne E&a. APTD-1563-Vol. 1 (PBZ39247/0), Exsrnfvt 

5- APTD-13W- Vo!. 2 (PB239248/8), Technical Sath; 

APTD-1SVol. 3 PWW9/6). Tcchnlcal Won Appndlas; 

F. SMaumm, APTDIJ64Vol. 4 (PB239W/4), Ea&e Dcslgn 

Opttmlrrrtloa. MoaJslrto Restar& Cap., St. lads, MO, June 

1973. 

13. DWmm, L.R.; DiB&, FA.; and Kopkw, M.D.: An RC-I 

OrganiSRnnktaeBoctomfng~for~~Mtwl~. 

(TE4322-251-83, Thmon Corp.; NASA fhrrtrad 

DEN3-302.) NASA CR-166256, 1W. 

14. Ww, Mlrhad D.; d DiNmm, b: Strtus Report on Dtisd 

orp& Ralkk compound E&nc far LDq Hrul Trucks. 

Arnomdtve Tc&adopg Wzbpmmd CorrtrMors'Coa.s 

hf&ng. Conl. 811090, Oa. 1P81, pp. 45-56. 
15. IQdfa, HE.: Waste Hcu Rmway fmm CLdir'LJtrc Dhul 

E&a by Exhaust-Driven byton Cpdes. (R63-95043, U.nhxl 

Tcchnobgk Rcoeprch Cema, NASA NASA DEN3-304.) 

NASA CR-168257, 19%. 

16. DhrdOgd~ RpnHnr+& Compound hgh far Long-H.lll 

Tmb. TE 4257-168-83, Tbenn0-w Cap., WatthPm, MA, 

Ang. 1983. 

17. Indm Trcnde and Stntktb. Canmm Carrta Jotrmsl, vol. 

140, M. 6, Jm 1983, pp. 99, lm, 102.10(,106. 

18. Fosta,RfchprdW.:~wnnlMukuR&cPrthonDksd~ 

oDdnPrsBDidTnrckswkh~RanklncB&tadngCycle 

Systenu (ORCS), Erhn-Fosta T dmdcrgy Asdates Inc., St. 

Jh,MI,Apr. lB1,pp. 27, Db. 

19. Smtvhorn, Larry W.; md W, Yiaor: Fuure Hcdny Duty 

T- Engh RtqutrcmcaU of thc mySc-ond Autmnodvc T- kdopaxt ContrPdm' r 

cocdbdm Mmlng. sm P-155, SAE, Wrrrendnk, PA, 1!w, 

pp. 2 2a 

20. 5'ktsm Wayne S.: Wpboord Enew Saw ~~II!I he kcu 

Sm. UW. Eng. J., vul. 94, m. 1, Feb. IWu, pp. 63-69. 

21. Enpl, P.K.; W-, R.; and -n, R.E.: Cumdm md I 

FOUL!! PI?& In Did Exhasu. J. Eng. Power, Vd. 101, m. I 

4, Oa. 1919, pp. JM. 

TZ Robats, P.B.; and MPrmn, H.D.: Soot and thc Combfncd Cyds I Boils. ASME Papa m47, Mar. 1979. 

23. memsl bdgadcm of Foulha md of a Wac 

Heat i-kcmuy Ha ExrfLanea. TE-4m-199-84, -illcmD 

Ekarm Corp., Waltham, MA, 198.5. 
NASA TM-86953 

2n.u ~fbcipblmf.~~ 

L~pd~ 

Cmparative Evaluation of Three A1 ternati ve Power 

Cycles for Waste Heat Recovery from the Exhaust of 

Adiabatic Dlesel Englnes 

7. -8) 

H. Murray Bailey 

a+- 

July 1985 

a-h-a 

778-34-1 2 

8.PrtamlnOWm-bnmNa 

E-2195 

1awurdI Fb 

aPahmdnammmmHanrrdAdrirar 

National Aeronautics and Space Administration 

Lewis Research Center 

Cleveland, Ohio 44135 

tzspamrln~~~arcy~anmdU.S. Department of Energy 

Office of Vehicle and Engine R&D 

Washington, DL. 20545 

1l.CartnclaCifudClo 

sp -- of ~aps -I; F,. w M 

1echn:ca 1 1', orandurn 

'lLspmofhg~~-- 

DOE/NASA/jOl94-43 

IS~Notsr 

Final report. Prepared under Interagency Agreement DE-AI01-80CS50194. 

1h - 

Three altemdtlve power cycles were canpared in application as an exhaust-gas 

heat-recovery system for use with advanced 'adiabatic" diesel engines. Tne 

power cycle a1 ternatlves considered were stem Rankine, organic Rankine wl th 

RC-I as the working fluid, and variations of an air Brayton cycle. The cumparison was made in tern of fuel economy and economic payback potentlal for heavyduty trucks operating In line-haul service. The results indicate that, In terms 

of engfne rated speclf i c fuel consumptlOn, a dsesel /a l ternatf ve-power-cyc le 

englne offers a significant Improvement over the turbocompound diesel used as 

the basellne for comparison. Tne maxlmurn improvement resulted frurn the use of a 

Ranklne cycle heat-recovery system in series with turbocompounding. The air 

Brayton cycle alternatives studied, which included both simple-cycle and compression-Intercooled conf lguratlons, were less effective and provlded about ha If 

the fuel consumption lrrrprovement of the Ranklne cycle alternatives under the 

same condltlons. Capltal and rnalntenance cost estimates were also developed for 

each of the heat-recovery power cycle systems. These costs were integrated with 

the fuel savlngs to identify the time required for net annual savings to pay 

back the Initial capltal Investment. The sensitivlty of capltal payback tlme to 

arbitrar Increase; in fuel prlce, not accompanied by corresponding hardware 

cost lnf 7 atlon, was also examined. The results indicate that a fuel prlce 

Increase is required for the alternative power cycles to pay back capital wltnin 

an acceptable tlme perlod. 

17.rc.r--bY-D 

Diesel technology 

Waste heat utllfzatlon 

Brayton cycle 

Rank i ne cyc 1 e 

MDbbCkRlon8ammTl 

Unclasslfied - unlimited 

STAR Category 85 

WE Category uC-96 

ia~cww.(ol~~1nparj 20.~~lrcn.ed1~1 

MCI assWa "*"?k" 22 Mu* Unclasslfted A02