doe/nasa/0361- 1 · available froli national technical information service u s department of...

8.

DOE/NASA/0361- 1 NASA CR- 1 80833 CTR 0723-S7001

Technical and Economic Study of Stirling

and Rankine Cycle Bottoming Systems for

Heavy Truck Diesel Engines

I. Kubo Cummins Engine Co., Inc.

Columbus, Indiana 47202

Prepared for NATIONAL AERONAUTICS AND SPACE ADMINISTRATION

Lewis Research Center

Cleveland, Ohio Under Contract DEN 3-361

for

U S . DEPARTMENT OF ENERGY

Conservation and Renewable Energy

Office of Vehicle and Engine R&D

Under Interagency Agreement DE-A1 01 -86CS50162 (hASA-CR-180833) T E C f i E l C A l dbI; € C C Y O I i I C N87-2E470

,C'IUDY OP S T X B L I B G AND f r A N K f b E CYCLE E C Z T C H I N G SYS'XEMS FOE E E A V Y ' I Z U C R DIESEL E b G I N E S E i n a l ReFcrt { C u n a i r E Enqiae Co.) U n c l a s 171 p Avail: N'IIYS HC B C 8 / E F A01 CSCL 13F G3/85 00975E4

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https://ntrs.nasa.gov/search.jsp?R=19870019037 2020-01-24T07:15:03+00:00Z

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1Codes are used for pricing all publications The code is determined by t h e number of pages in the publication information pertaining to the pricing codes can be found in the current issues of the following publications which are generally available in most libraries Energy Research ASc',ac'-c fEf iA1 G o J e r n q e i : Reporfs Annodncernents and Index (GRA and I) S c j e i f i f c and Technics' Absrrac' Reports (STAR) and publication, NTIS-PR-360 available from NTlS at the above address

DOE/NASA/0361-1 NASA CR-180833 CTR 0723-87001

Technical and Economic Study of Stirling and Rankine Cycle Bottoming Systems for Heavy Truck Diesel Engines

1. Kubo Cummins Engine Co., inc. Columbus, Indiana 47202

1887

Prepared for NATIONAL AE RONAUTlCS AND SPACE ADMlNlSTRATlON Lewis Research Center Cleveland, Ohio Under Contract DEN 3-361

for

U.S. DEPARTMENT F ENERl

Conservation and Renewable Energy Office of Vehicle and Engine R%D Under Interagency Agreement DE-A1 01 -86CS50162

FORWARD

The author would like to thank the guidance as well as helpful inputs by Murray Bailey of NASA Lewis during the execution of the entire program.

theif effort as subcontractors are recognized. Following companies participated in this program and

- Mechanical Technology Inc., D.

- Argonne National Laboratory, R. R.

- Adiabatics Incorporated, R. - CMC Aktiebolag, S.

In addition, the author would like cooperation by Foster-Miller, Inc. and Corporation which performed conceptual

Najewicz Kamo Carlqvist Sekar Cole

to acknowledge the Thermo Electron designs of Rankine

bot%oming systems under a previous program for NASA Lewis. Their cooperation were essential for us to complete the cost estimates of both Rankine systems.

In Cummins, many people participated in the program. Following are members of the bottoming cycle study team:

W. Brighton - Design J. Wagner - Engine Application/Service R. Linney - Manufacturing Cost Estimates D. Reese - Thermal Analysis/Design

PRECEDING PAGE BLANK MOT FiLMEo

iii

CONTENTS

I . INTRODUCTION . . . . . . . . . . . . . . . . . . . . . .

c

I1 . DIESEL STIRLING SYSTEM STUDY . . . . . . . . . . . . . . 1 . "STATE OF THE ART" TECHNOLOGY . . . . . . . . . . . .

1.1 SCREENING OF COMPOUND ENGINE CONFIGURATIONS . . 1.1.1 RESD V-4 Engine Concept . . . . . . . . . 1.1.2 Single-Acting V-4 (SAV-4) Engine Concept

1.1.3 FPSE/Hydraulic Transmission . . . . . . . 1.1.4 Summary of Concept Evaluations . . . . .

1.2 Detailed Analysis/Conceptual Design . . . . . . 1.2.1 Engine Size . . . . . . . . . . . . . . . . 1.2.2 Single-Stage vs . Double-Stage

Power Recovery . . . . . . . . . . . . . 1.2.3 Heater Head Redesign . . . . . . . . . . 1.2.4 Other Engine Modifications . . . . . . . 1.2.5 Summary of Conceptual Design/Analysis . .

1.3 MANUFACTURING/MAINTENANCE COST ESTIMATES . . . . 1.3.1Manufacturing Cost . . . . . . . . . . . 1.3.2 Maintenance Costs . . . . . . . . . . . .

1.4 CONCLUSIONS FOR THE STIRLING SYSTEM WORK . . . . 2 . HIGH TEMPERATURE COMBINED DIESEL/STIRLING CYCLE . .

2.1 Preliminary Arralysis . . . . . . . . . . . . . . 2.2 Conceptual Design . . . . . . . . . . . . . . . 2.3 Computer Sinulation . . . . . . . . . . . . . . 2.4 Conclusions on the TSA-cycle Study . . . . . . .

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Page

. . . 111. COMPARATIVE EVALUATION OF THREE BOTTOMING CYCLES. 14

3.1 FUEL ECONOMY EVALUATION . . . . . . . . . . . . . . 3.2 MANUFACTURING COST ESTIMATES. . . . . . . . . . . . 3.3 SERVICE/MAINTENANCE COSTS . . . . . . . . . . . . .

14

15

16

3.4 ECONOMIC EVALUATION . . . . . . . . . . . . . . . . 17

. . . . . . . 3.5 CONCLUSIONS OF COMPARATIVE EVALUATION 18

IV. INTEGRATED DIESEL/RANKINE SYSTEM. . . . . . . . . . . . 19

. . . . . . . 4.1 THERMAL ANALYSIS OF THE ENGINE SYSTEM 19

4.2 FLUID SELECTION AND OPTIMIZTION OF RANKINE CYCLE. . 20

4.3 CONCEPTUAL DESIGN OF THE SYSTEM 21 . . . . . . . . . . . . . . . . . . . . . . 4.3.1 Heat Exchanger Sizing 21

. . . . . . . . . . . . . . . 4.3.2 Power Expander. 22

4.3.3 Engine Design/Layout. 23

4.4 PERFORMANCE PREDICTION. . . . . . . . . . . . . . . 23

. . . . . . . . . . . .

4.5 MANUFACTURING COST ESTIMATE . . . . . . . . . . . . 4.6 ECONOMIC EVALUATION . . . . . . . . . . . . . . . .

23

24

V. CONCLUSION. . . . . . . . . . . . . . . . . . . . . . . 25

. . . . . . . . . . . . . . . . . . . . . . . . . . . 27 FIGURES

TABLES. . . . . . . . . . . . . . . . . . . . . . . . . . . . 65

APPENDIX 1 - M.T.I. COST STUDY OF STIRLING BOTTOMING . . . . . . . . . . . . . . . . . CYCLE SYSTEM. 86

APPENDIX 2 - THERM0 ELECTRON CORP. COST STUDY OF ORGANIC RANKINE BOTTOM CYCLCE SYSTEM . . . . . . . . . . 103

APPENDIX 3 - FOSTER - MILLER INC. COST STUDY OF STEAM BOTTOMING CYCLE SYSTEM . . . . . . . . . . . . . 113

APPENDIX 4 - INTEGRATED BOTTOMING CYCLE FOR TRUCK DIESEL ENGINES. . . . . . . . . . . . . . . . . . . . . 125

APPENDIX 5 - ADDITIONAL COST FOR INTEGRAL BOTTOMING CYCLE ENGINE V8. L10 TURBOCHARGED AFTERCOOLED ENGINE . 165 -

vi

SUMMARY

The main objective of the program is to evaluate the concept of bottoming cycle to heavy duty transport diesel engine applications. For that objective, following sub-objectives are set under this program.

I - Develop conceptual design and cost data for a Stirling bottoming cycle system,

- Life-cycle cost evaluations of three bottoming systems: Organic Rankine, Steam Rankine, and Stirling cycles.

research . - Suggest future directions in waste heat utilization For the Stirling bottoming cycle, MTI, Albany, N . Y .

completed the study under the ground rule that the system only utilized "state of the art" technology. conceptual design of a system which would fit in a "cab-over" type truck was developed and manufacturing cost estimates for the system were performed.

Adiabatics Inc. and Stig Carlqvist of CMC Aktiebolag studied a new Stirling system called "high temperature combined cycle@@. Results of the "high temperature combined cycle" indicated that its combined thermal efficiency could reach 519, similar to the diesel engine with a Rankine bottoming cycle. However, the system requires extremely high temperature materials and lubricant which are well be ond the current technology.

Therefore, the exhaust-gas heat into the intake air. improvement is considered too small for the risk factors involved in developing the system.

by Eflll:, completed the conceptual design phase of the bottoming cycle evaluation based on the life-cycle cost analysis. results of the studies made by Foster-Miller and Thermo Electron respectively under a previous DOE/NASA program were used. Variables considered are initial capital investments, fuel savings, depreciation tax benefits, salvage values, and service/maintenance costs. Fuel savings are based on the truck mileaye improvements calculated with the Cummins VMS (Vehicle Mission Simulation) computer code. All bottoming systems are to be used with advanced "adiabatic" engines. Comparisons were made against a turbocompound engine, seriously considered as a way to improve fuel economy for heavy duty

A

In addition to the above work on the Stirling system,

Furthermore, a regenerator P s required to recycle

The above Stirling system study, particularly the work

For steam- and organic-Rankine systems,

truck applications.

provide a 18 to 19% IFtR (Internal Rate of Return) investment opportunity for truck owners. However, currently none of the three bottoming systems studied are even marginally attractive. Manufacturing costs of the systems have to be reduced by at least 65% in order to become competitive against the TCPD engine. A new innovative approach is required for any bottoming system to be applied for heavy duty truck engines.

As such a system, an integrated Rankine/Diesel system was proposed. cylinders as an expander. This would eliminate the need for the power transmission devices required for all conventional bottoming systems. Control requirements would be less. Another aspect of the proposed system is the capitalization of in-cylinder heat loss which is quite substantial for the I8adiabaticn engine. The concept reduces the size of the exhaust evaporator.

Conceptual design of the system and a rough economic evaluation were made. potential to become an attractive package for end-users, giving approximately a 20% IRR at the fuel cost of $1.25/gallon.

The turbocompound/aftercooled (TCPD) engine would

The system utilizes one of diesel

Results indicate the system has a

The study was intended for a rough evaluation of the concept and optimization of the system was not performed. Further optimization is possible by eliminating/combining some of the concepts built in the current design.

2

I . INTRODUCTION The main objective of this waste heat utilization

study is to evaluate the economic feasibility of the bottoming cycle concept to heavy duty truck engine applications. candidates which provide a good fuel economy improvement and should be evaluated under the study. Rankine, Brayton, and Stirling cycles fit the qualification.

In 1985, M. M. Bailey of NASA Lewis reported a comparative evaluation of three alternative bottoming power cycles (ref. 1). Alternatives studied were steam Rankine, organic Rankine, and an air Brayton cycles. The study was made under the following ground rules:

- Base engine is an "Adiabatic" turbocharged- diesel. - Engine output is 350 HP. - Concepts for bottoming cycle systems only use "state of the art" technology (1985 - 1987 time period) .

There are several bottoming cycle

Currently,

Results indicated that the Rankine cycles were substantially better than the Brayton cycle in terms of the payback to truck owners. included due to the lack of a conceptual design of the system at that time.

The Stirling cycle was not

Under this program of which the major object is to complete the study initiated by Bailey, following tasks were set:

1. 'Development of conceptual design and cost data for diesel/stirling system,

2. Life-cycle cost evaluation of three bottoming systems: Organic Rankine, Steam Rankine, and Stirling cycle.

3. Preliminary evaluation of a new integrated Diesel/Rankine system.

Under the task 1, two concepts were evaluated. One is to follow the ground rules outlined above. This concept would complete the Bailey's comparison study. The other concept is to use much more advanced technologies including new materials and lubricants. Therefore it can not be compared to other bottoming systems being evaluated here. However, this study was included because it was believed that a new approach to the bottoming cycle concept would be needed to make it economically competitive against the turbocompound engine.

3

The task 3 was an addition to the original program. Based on the evaluation study of task 2, a new concept to integrate a diesel engine with a Rankine bottoming system emerged. Here, a preliminary conceptual design/analysis was made on the system.

11. DIESEL STIRLING SYSTEM STUDY

1. "STATE OF THE ART" TECHNOLOGY

This portion of the study was performed by Mechanical Technology Incorporated, Latham, N.Y. Details of the study is described in a MTI report "MTI 87SESD33" (Ref. 2 )

1.1 SCREENING OF COMPOUND ENGINE CONFIGURATIONS

There are many types of mechanical drive arrangements for a Stirling engine such as crank/connecting rod, rhombic drive, wobble plate, and free piston/hydraulic converter. based on complexity, durability, performance, size, and weight; three basic configurations are selected for further examination. crosshead concept, which is represented by the RESD V-4 (MTI automotive Stirling engine), a double-acting, four-cylinder l@V1v configuration; the SAV-4, a single-actin?, pressurized crankcase concept: and a FPSE/hydraulic converter concept.

1.1.1. RESD V-4 Engine Concept

After a brief review of the arrangements

They are a connecting rod with

A cross section of the RESD V-4 is shown in Figure 2-1. Coolers and regenerators in this engine are arranged in an annular configuration about the piston, thus minimizing the number of pressurized parts in the engine, particularly the heater head castings. The working gas is sealed from ambient pressure at the piston rod seal. The rod seal utilizes a type of sliding seal known a8 a pum ing Lenningrader (PL) seal. Main

the connecting rod/crosshead assembly is of conventional design. The engine is based on the technology developed under the Automotive Stirling Engine (ME) program and is supported by eight years of testing on Mod I and P-40 Stirling engines.

bearing8 are o H 1-lubricated rolling element bearings, and

It is the most compact engine of the three concepts,

4

and has advantages of smooth torque output and good manufacturability. associated with the life of the piston rings and sliding rod seals. Currently, the design life for these items is around 3,500 - 5,000 hours.

A parametric study on the performance of the engine was performed to understand the effect of Stirling exhaust temperature and engine speed. the power recovery maximizes in the exhaust temperature range of 700-900°F (i.e., heater head temperature of 600-80WF) for 2,000 rpm Stirling engine speed. Considering the fact that the size and weight of the engine increase with reducing exhaust gas temperature, the optimum temperature is determined at around 80eF. It would give a Stirling power recovery of 30-31 hp with a turbocharged engine as a base.

Disadvantages of the engine are

As shown in Figure 2-2,

1.1.2. Single-Acting V-4 (SAV-4) Engine Concept

The concept uses single-acting pistons as shown in Figure 2-3. Two cylinders are used as compressors and - the other two act as expanders. Therefore, during one revolution of the crank, two power strokes take place, unlike four power strokes for the double acting four cylinder engines. sliding seals, this concept utilizes a pressurized crankcase with rotary oil lubricated seals for sealing the crankshaft. Piston rings seal between the cycle pressure variation and the cycle mean pressure in the crank case. Connecting ducts between the top of pistons and the bottom of other pistons used for the double-acting engine are eliminated. This would minimize the dead volume and improve the engine performance.

A dry-lubricated design is required between piston/ cylinder liner. Sealed grease-lubricated, rolling element bearings are used at both ends of connecting rods. The oil lubricated face seals are located outside of the en ine and isolated from the crankcase by "lip seals"

The advanta ea of the SAV-4 engine include the overall simplic 1 ty of the design, a heater head configuration well suited for the bottoming cycle application, and a seal location easily accessible for maintenance.

as well. Results are similar to the RESD engine, except an improvement in power recovery (10% v. 11.5% in BSFC improvement at the rated condition) . The parametric performance curves are similar to the RESD results. However, the SAV-4 is significantly heavier than the RESD V-4, mainly due to the inherent nature of the single- acting v. double-acting concept.

In an attempt to eliminate the use of

wh s ch do not experience any pressure differential.

A performance analysis was made for the SAV-4 engine

5

. There are several technological Unknowns associated

with this concept.- The life of the rotary seal is not well understood as is the problem of hydrogen diffusion through the lubricating oil. concept is also a major source of uncertainty.

The dry-lubricated piston

1.1.3. FPSE/Hydraulic Transmission

The engine is similar to that shown in Figure 2-4, except that no combustor is required and that the linear alternator in the figure would be replaced with a hydraulic transmission. The engine would use a tubular heater head with an annular regenerator and an cooler. The hydraulic transmission utilize metal bellows or a metal diaphragm to pump a hydraulic fluid. The high-pressure fluid then drives a hydraulic motor to produce a shaft power. The advantage of this approach is that the Stirling engine is essentially decoupled from the diesel operation, i.e., Stirling engine speed is independent from the diesel speed and its location is not restricted by the location of the diesel crankshaft. The engine is hermetically sealed and usage of any sliding or rotating seals are minimized. However, the use of the hydraulic transmission/motor significantly increases the complexity of the system since a hydraulic circuit would require valves, accumulators, and many more components.

with regard to long life and flexible operation; however, the state of development of the engine is several years behind that of kinematic Stirling engines, and the size required for this application, 25-30 KW, is clearly larger than any existing FPSE under development today.

are shown in Figure 2-5. speed of 2,500 r p m to be compatible with its hydraulic system and the heater head temperature was set at 700°F based on the MTI's previous experience. Efficiencies for hydraulic pump and motor were assumed to be 97 and 95% respectively. BSFC improvement of 8.58% was achieved with the system.

Overall, the FPSE had several potential advantages

Performance of the FPSE was analyzed and the results It was optimized at the engine

1.1.4. Summary of Concept Evaluations

In addition to the performance discussed in the previous section, maintenance and manufacturing costs for each system were estimated. The results are shown in Tables 2-l through 2-3. three systems is summarized in Table 2-4 in terms of several factors as performance, maintenance and manufacturing costs, and size. It appears that the RESD

The overall comparison of the

6

v-4 is the best choice for bottoming cycle applications based on technical maturity, performance, and packagiability.

1.2 Detailed Analysis/Conceptual Design

After the basic configuration was selected, the integration of the system with the diesel was evaluated more in detail. The space available for packaqing the Stirling bottoming system was one of the first issues addressed. performed next. The optimization included the comparison of single-stage V. double-stage arrangement as well as redesigning of the heater head to improve performance. A brief summary of the work is described below.

Heater head design optimization was

1.2.1 Engine Size

The available packaging space around the diesel in a heavy-duty truck is very limited. Using the Iwcab-over" truck configuration in mind, the space was determined and is shown in Figure 2-6. preliminary screening study was found to be significantly larger than the space available for the system. The MTI Mod I1 engine being developed for the automotive application, on the other hand, is smaller than the RESD V-4 engine (approx. 30%) and fits in the space. Therefore, it was decided to use the M o d I1 ASE engine for this program. bottoming cycle based on the ASE is that the use of the existing technology would minimize the cost of development as well as any technical barriers for the introduction of the engine.

The RESD engine used for the

A benefit of development a Stirling

1.2.2 Single-Stage v. Double-Stage Power Recovery

Performance was analyzed using both a single-stage operation and a double-stage operation over the Stirling engine speed range. transfer calculations between the diesel exhaust gas and the heater head metal as well as that between the heater head metal and the working fluid. coefficients used for the analysis are based on data developed by MTI during rig testing of heater head segments (Ref. 3). Results are shown in Figure 2-7.

As shown in the Figure, performance difference is not significant for this particular size engine and operating conditions. Therefore, a single-stage operation was considered for this study. improvement of the Mod I1 size engine was found to be 7.25%.

The analysis involved detailed heat

Heat transfer

The net performance

7

1.2.3 Heater Head Redesign

In order to improve the performance of this bottoming system, following engine modifications were considered: simplification of the heater head design by incorporating a individually tubed heater head instead of the four manifold arms used in the ASE engine, and increase in the heater tube length.

Elimination of the complex heater heat castings would provide: 1. cost reduction, and 2. the cycle performance improvement due to a reduction in the dead volume. The effect of the increase in heater tube length is significant as shown in Figure 2-8. However, if the lenqth is increased more than a twice of the current design, flow losses in the heater tubes would negate the improvement and output would be reduced. head design is shown in Figure 2-9.

arrayed around each heater head. The tubes are finned only on the rear row and the fin spacing is a 0.5 mm (0.022 in.). The detailed heat transfer in the heater head area is studied further by using a two-dimensional computer program for a single-row tube heat exchanger with plate fins (MTI's FIN2D program). The analysis indicates the tube configuration described above will provide a heat transfer capacity of well above 80 KW required at the design point for this application.

1.2.4 Other Engine Modifications

The new heater

Each heater head will utilize 30 U-shaped heater tubes

The drive system will be essentially the same as the Mod I1 except that the hydrogen compressor will not be required for this application, since the engine will operate at a constant pressure. However, because of the leakage associated with the static and sliding seals, a hydrogen makeup tank will be required. A four-liter hydrogen storage bottle will be used and be recharged every six months. drive system is shown in Figure 2-10.

The overall mounting of the Stirling engine, including the power coupling is shown in Figures 2-11 and 2-12. As shown, the flywheel housing of the diesel is moved aft by approximately 1.5 inches to add an intermediate spacer/ chain cover. The Stirling engine itself is mounted on two brackets, one of which is supported by the flywheel housing and the other by the transmission casing. The power coupling incorporates a simple direct chain drive with torsional isolators and an electric clutch. The chain is a commercially available 1.5-in. nsilentn chain. Since the rated speed of the diesel is 1,900 r p m , the drive ratio is nearly 1 to 1 (20:19). The electric

A cross section of the cold engine

8

clutch will allow the Stirling engine to be disconnected either for heater head cleaning or during starting or extended idling of-the diesel. disconnection of the system in the event of a system malfunction. thermocouple on the Stirling heater head. temperatures below the Stirling self-sustaining point ('3500F), the Stirling would not create a parasitic loss on the diesel.

It also allows

The clutch would be actuated by a Thus at

Control of the Stirling bottoming system is simpler than the Stirling engine control required for the independent system. fixed level set by a pressure regulator on the h drogen makeup tank. During a @@down-throttle" of the d x esel, a "short- circuit" valve on the Stirling would connect the engine cycles such that the pressure-waves in cycles 180° out of phase cancel, reducing the output power instantaneously.

heat at the desiqn point. Thus it requires a water pump and a radiator similar to a conventional automotive one. The pressure drop of the diesel exhaust gas through the system is estimated to be on the order of 10 to 12 inches of water.

The engine pressure will remain at a

The Stirlinv engine will reject approximately 53 KW of

1.2.5 Summary of Conceptual Design/Analysis

The Mod I1 design was chosen as a final configuration for this study. The system will fit in the envelop available in a cab-over truck. under various engine operating conditions are shown in Table 2-5. At the diesel engine rated condition, it improves the BSFC of 9% for a turbochar ed engine. Figure 2-13 shows BSFC improvement v. d P esel engine load for 1300- and 1900-rpm operations.

Performance of the system

I - I

1 . 3 MANUFACTURING/MAINTENANCE COST ESTIMATES

1.3.1 Manufacturing cost

MTI developed cost estimates for the bottoming cycle

year. It utilized the estimate prepared by the P I;er oneer system based on a production rate of 10,000 units

Engineering and Manufacturing Co. for the Mod I1 ASE engine. The original MTI estimate was $1,789 which represents a "matured" manufacturing costs (ref. 2). Cummins estimated the cost to be at $2,281 which represents a 70% of the estimated manufacturing costs based on the M o d I1 drawings (See Appendix 1) and is supposed to be the matured costs.

9

The difference between the two estimates are due to; 1. labor rates used, 2. the manufacturing method assumed to be used for the system, and 3. quality of components used for the system. Cummins used a labor rate close to a $70/hour, while MTI assumed a rate of $30 to $35/hour. As for the manufacturin method, Pioneer's estimate is

may fie justifiable, if the Mod I1 ASE engine is a reality. The Cummins estimate, on the other hand, is based on a much less capital intensive method of manufacturing. Since Pioneer's estimate is for the automotive application, the components are all automotive quality parts. However, items such as the radiator required for our application are quite different from the automotive application. Therefore, Cummins' estimate was based on components of industrial quality.

Since all the bottoming cycle systems studied under this program are evaluated by the Cummins approach and also the ASE engine is still quite uncertain, the estimate made by Cummins will be used for the economic comparison of those bottoming systems.

1.3.2 Maintenance Costs

based on the use of cap 4 tal intensive processes. This

MTI estimated service/maintenance costs based on scheduled periodic overhauls, yearly general maintenance, and operator capital costs. The scheduled periodic maintenance is associated with major engine overhaul that will be required at 5,000-hour intervals for seal/ring replacement, as well as replacement of bearings and other renewable items. During the 7-year period (14,000 hours), two engine overhauls will be required and its yearly cost was estimated to be at $135.63 per year.

such items as oil change, cooling fluid replacement, hydrogen recharging and other general maintenance. was estimated to be 5% of the retail engine cost per year and, based on the MTI price, is $179/year. (Based on the Cummins cost ($4 ,562) , it is $228/year) The last item, operator's capital, represents the capital investment by fleet operators to install support equipment and to train personnel. It is based on a 25-vehicle fleet and a 5-year recovery of the capital investment. The result was $116/year.

cost and $480/year with Cummins cost. however, are based on the MTI assumptions. Cummins also made the estimate and obtained different numbers as shown in the later chapter of this report.

The yearly general maintenance of the engines cover

It

Total service/maintenance cost became $431/year by MTI Both figures,

10

1.4 CONCLUSIONS FOR THE STIRLING SYSTEM WORK

A conceptual design for a Stirling bottoming cycle system has been performed. The engine meets the packaging requirements of the heavy duty truck application and provides a 9% fuel savings over a baseline turbocharged adiabatic engine. No serious technical barriers can be foreSeen for the system, though it still requires a limited amount of development in the areas such as heater tube fouling/cleaning, development of long-life seals/piston rings, and confirmation of the control approach.

2. HIGH TEMPERATURE COMBINED DIESEL/STIRLING CYCLE

This portion of the study was performed by the

i) Preliminary Analysis ----.. CMC Aktiebolag

ii) Conceptual Design ------ CMC Aktiebolag iii) Computer Cycle Analysis ------ Adiabatics Inc.

following subcontractors: i

Details of the study are described in a final report by Adiabatic Inc. (ref. 4)

A schematic of the concept is shown in Figure 2-14. Since the high exhaust gas temperature is critical for the performance improvement of any bottoming systems, the concept utilizes a heat recirculation from exhaust gas to intake air so as to obtain a extremely high exhaust gas temperature. A T-S diagram of the cycle is shown in Figure 2-15.

2.1 Preliminary Analysis

A preliminary thermodynamic calculation was made by using the T-S diagram shown in Figure 2-15. figure, combustion process is divided into three parts, C1, C2, and C3 which represent constant volume, isothermal and decreasing temperature expansions respectively. Expansion process is also separated into three parts as isothermal, decreasing temperature, and adiabatic expansion processes. STH is the part where the heat is extracted into the Stirling bottoming system.

In the

Some of the assumptions made for the study are:

11

- In-cylinder conduction heat loss is assumed to be 5-89

- Stirling engine efficiency is 40%. - Mechanical efficiency of the diesel engine is 90%. - Turbocharger overall efficiency is assumed to be in a range of 65-66.69.

- Combustion as dissociation effect is not taken into the considerat s on. (Variations in Cp/W with temperature and air/fuel ratio are assumed to be linear.)

of total fuel energy.

- Peak cylinder pressure limit was set at 2,500 psi. The results of the calculation by S. Carlqvist are

shown in Figure 2-16 and table 2-6. It indicates that the total system efficiency obtainable with the concept is 60.52. However, as shown in the table, there is a large error i n the heat balance between the heat input to the engine and the total heat out from the engine including the diesel shaft output and the heat to the stirling system (Total output is 24% hi her than the heat input, 602.7 vs 747.4.). used for Cp during the c cle calculation. Therefore, the final result is not cons x dered to be accurate. This is pr 9 marily due to wrong values

If, instead of increasing the heat input as done by S. Carlqvist, the output from the diesel was decreased in order to balance the energy, resulting thermal efficiency becomes 53.6%, much smaller than the 60% shown in the figure and closer to the computer simulation result described in the next section.

2 . 2 Conceptual Design Conceptual designs at several levels of integration

were proposed. Unlike the MTI Stirling design, there was no space limitations imposed for this study. 2-17 and 2-18 show the layouts of semi-integrated systems, while Figure 2-19 depicts a fully-integrated system.

Figure8

2.3 Computer. Simulation

Performance evaluation of the TSA-c cle was made with a use of a diesel cycle simulator by Ad x abatic Inc. order to calibrate/evaluate the simulation program, calculations were made on a Cummins L-10 turbocharged aftercooled engine case with the simulator and remultm were compared against actual engine data obtained by Cummins. After the accuracy of the computation was assured, the TSA-cycle calculation was made. Results of the calculation are shown in Table 2-7 and Figure 2-20.

In

12

The overall efficiency of the TSA-cycle is 50.7%. Since the base turbocharged "adiabatict1 engine has an efficiency of 432,-the fuel econom improvement with the

Rankine cycle which shows the 16% BSFC improvement. Combination of the TSA-cycle with a turbocompound engine was also evaluated. However, the thermal efficiency did not improve.

TSA-cycle is 17 to 182, slightly h x gher than the steam

2.4 Conclusions on the TSA-cycle Study

As mentioned before, the fuel economy improvement with the system over the baseline engine was only slightly better than the steam Rankine cycle. However, there are many technical obstacles, such as high temperature materials/lubricants, to be overcome before this system can be produced. The material temperature is extremely high, even way above some of the ceramic material- capabilities. Therefore, it was concluded that the s stem was not worth ursuing any further, unless much h x gher pressure capab f lities can be established such that the efficiency improvement can become significantly higher than the conventional bottoming systems.

13

111. COMPARATIVE EVALUATION OF THREE BOTTOMING CYCLES

With the inforiation on the Stirlin bottoming cycle available, the task of comparing bottom 9 ng cycles was initiated. The cycles evaluated and subcontractors who performed the conceptual design/performance analysis for each system are as follows:

- Stirling cycle ... MTI - Organic Rankine cycle ... TECO (Ref. 5) - Steam Rankine cycle ... Foster-Miller (Ref. 6) The analysis was made based on a life cycle costs/

benefits to end-users. Variables considered for the analysis are:

- Initial Capital Investment - Future Incomes/Expenses . Income: Fuel Savings

Depreciation Tax Benefits Salvage Value of the engine

. Expenses: Service/Maintenance Costs

Operational/economic assumptions made for this analysis are tabulated in Table 3-1. As shown in the table, sensitivity analyses were made for different oil prices and base truck fuel mileage.

3.1 FUEL ECONOMY EVALUATION

Instead of using the steady state rated condition to represent the fuel economy improvement, the Cummins Vehicle Mission Simulation (VMS) program was used to evaluate the improvement in truck mileage with bottoming cycles. Three routes are considered for the study. Those are: Reno-Sacramento, Indianapolis-Chicago, and Columbus-Louisville-Cincinnati-Columbus. The first one represents the most hill case and the second is one of the flattest route in th x s country. or less a mixture of hills and flat routes and is considered as a standard route which represents a typical truck route in this country.

procedure was used:

each of the three cases. percentages of time spent for different engine operating

The last one is more

For obtaining the truck mileage improvement, following

- First, the VMS simulations were run one time for The results would give

14

conditions for the three routes.

- Engine performance maps for different engine configurations were obtained b simulation rogram and by pred I cting power recovery through var ! ous bottoming systems for different exhaust gas conditions.

the truck mileage would be estimated by combining the maps and the VMS results.

using the diesel cycle

-, The maps were divided into several areas such that

As shown in Table 3-2, the organic Rankine system gives the best improvement, while the Stirling system is the worst. The performance improvement is better on the hilly route than on the flat one. study, the results with the "mix" route were used. The reason for the poor performance of the Stirling system is illustrated in Table 3-3. The system relies on the high heater-head temperature to obtain a high power conversion efficiency. however, high heater-head temperature means high temperature of the Stirling exhaust gas which, in turn, reduces the efficiency of the energy extraction from the diesel exhaust gas.

various bottoming systems.

For our comparison

Because of its inherent characteristic,

Figure 3-1 shows the comparison in performance for

3.2 MANUFACTURING COST ESTIMATES

Since actual hardware for the or anic Rankine system were available at TECO from the prev 7 ous DOE program, the

Observation of actual hardware and draw 1 ngs were manufacturing cost estimates were made on that s stem first. utilized for the evaluation and detailed results are shown in Appendix 2. For the steam system, the estimate was made by comparing components with the organic system. Most of components are similar between the two systems, except the organic cycle uses a turbine type of expander with a rated speed of approximately 20,000 rpm. The steam system uees a two-cylinder reciprocating expander with a rated speed of 2,000 rpm. steam system is seen in Appendix 3.

Detailed study for the

Table 3-4 compares manufacturing costs for the three systems. Estimates by subcontractors were used for vapor generators of Rankine systems. The vapor generator of the organic system is more expensive than that of steam system due to mainly a by-pass mechanism required for the organic cycle to prevent over-heating of the fluid. The reciprocator type steam expander costs more than the turbine expander used for the organic s stem. In summary, organic system's cost is the h x ghest at $4,938

15

and steam Rankine costs second at $4,199. c cle s stem cost is the lowest of all at $3,258. figure 1s quite different from the one made by MTI as mentioned n a previous section. Figures for Rankine systems presented here, however, turned out to be close to those made by subcontractors in 1983.

All the cost figures are estimated from the current designs for each system. Considering the effect of the "learning curven, the costs were reduced by 309 to project "maturedn costs of the systems. Figures for the matured costs are shown in the parentheses. The "maturedn costs were used for the economic analysis in Section 3.4.

The Stirling The

Output levels differ among engines with the three different bottoming systems due to the difference in power recovery. Therefore a correction was made on the manufacturing costs for a same output level (@350 HP) based on the @f0.7-power law" correlation developed by Bailey for his analysis.

Actual prices which end-users have to pay for the bottoming cycles include mark-up b makers and OEMs. The mark-up changes factors including the entire engines. Therefore it is number for the mark-up value. Based on inputs from Cummins marketing area and also from a OEM, the total mark-up was set to be at 100%. costs were multiplied by 2 to obtain final prices for the end-users.

Thus the manufacturing

Final price figures are listed in Table 3-5. in the table, difference in price among the bottoming systems becomes much smaller by the constant output comparison.

As seen

3.3 SERVICE/MAINTENANCE COSTS

Two different sources were used to determine senrice/ a ~ummins 7-year maintenance/

Based on the data, repair contract and serv ce data accumulated for routine maintenance costs,

servicing/maintenance costs. following estimates are made:

namelI - Regular Semi-Annual Inspection & Service

- Other Services - Variable Element of Maintenance/Service

39 of Equipment Original Price/100,000 miles

7% of Equipment Price/100,000 miles

16

(Major overhaul and turbo replacement, etc . ) 50% of Equipment Price/500,000 miles

Therefore, the total annual service/maintenance cost was assumed to be a 20% of original equipment price.

3.4 ECONOMIC EVALUATION

Based on the various economic data generated above, a comparative evaluation of the systems was made by using IRR (Internal Rate of Return) as a measuring criterion.

A summary of the manufacturin

the estimates made by each subcontractor for its respective bottoming cycle system.

Results are shown in Figures 3-2 through 3-5.

and maintenance costs to be used for this study is shown P n Table 3-6, along.with

In Figure 3-2, IRRs for systems including a

ne are shown with fuel price

As seen n the figure, less than 102 while that

for the turbocompound engine shows a above 20% return with a fuel price of $l.OO/gallon. Steam Rankine system. But it, too, won't be attractive unless fuel price goes over $1.50/gallon level. The result is based on'the truck fuel economy of 8 miles/gallon. As shown in Figure 3-4, the net present value of the steam Rankine system becomes positive at a above 15% cost of ca ita1 level if the base truck fuel economy is below 6 m !! les/gallon. The better the base truck fuel economy is, the harder it is for any bottoming cycle systems to be used commercially. Due to future improvement on the base truck designs, the 8 miles/gallon assumption for the fuel economy is reasonable.

Figures 3-3 and 3-5 show similar comparisons as Figures 3-2 and 3-4, with a turbocompound engine rather than a turbocharged engine as a base.

either a turbocompound engine or another engine with a bottoming cycle system. From the figures, it is clear that the turbocompound engine is well superior to bottoming cycle systems unless fuel prices start to soar above $1.75/gallon levels. And even at a $2.00/gallon level, only the steam Rankine system becomes attractive and Stirling and Organic Rankine systems won't be feasible.

s" for this fi re is a

The best system is the

The will show us a comparison of systems when we have an opt x on to buy

As a final analysis, calculations were made to obtain a system price of each bottoming system which would be competitive against the turbocompound engine. The

17

results are shown in Table 3-7. For the study, fuel economy improvement with each bottoming system is assumed to be the same as the current system. A 8 shown in the table, more than 355 cost reduction is required for the steam system to be competitive to the turbocompound system. reductions.

Other systems require much higher cost

3.5 CONCLUSIONS OF COMPARATIVE EVALUATION

Based on the above analysis, following conclusions are made:

- Using a life-cycle costs/benefits concept model, an economic evaluation was made on three bottoming cycle systems and on turbocompound system.

- Assuming that fuel economy improvement with turbocompound engine (TCPD) is 6% over the baseline turbocharged/aftercooled engine and fuel cost is $l.OO/gallon, TCPD system would provide a above 20% IRR investment opportunity. However, the system still required more than two years of payback period.

- None of the three bottoming cycle systems are even marginall attractive, unless diesel fuel price becomes close to I 1 . 75/gallon.

- Manufacturing costs for bottoming systems have to be reduced, at least 35%, in order for them to become competitive against the TCPD engine in terms of return in investment .

Above results indicate that bottomin cycle systems as they are designed now will not be econom f cally feasible As a new approach of making a bottoming system attract x ve, for a foreseeable future, unless a totally different approach is introduced to reduce its cost substantiall . an integrated Rankine/diesel system concept has emerged. The concept w i l l be analyzed and evaluated in the next chapter .

18

I ' IV. INTEGRATED DIESEL/RANKINE SYSTEM

A schematic of the new inteqrated diesel/rankine cycle is shown in Figure 4-1. It utilizes one fluid for engine cooling as well as for a Rankine c cle. Thus the need for an additional radiator is elim x nated. In addition, the size of the vapor generator required in the exhaust system can be reduced due to the fact that the working fluid picks up heat energy through the engine cooling before it reaches the vapor generator. This is am important factor from the cost point, since the vapor generator cost is one of the major items of the total manufacturing cost.

In this chapter, a thermal analysis of the diesel/Rankine system is described. Based on the available heat, a selection of a working fluid and a Rankine cycle optimization were made. Finally, a conceptual design of the system was proposed and a economical evaluation of the system was made.

4.1 THERMAL ANALYSIS OF THE ENGINE SYSTEM

Figure 4-2 shows passages for the working fluid. As shown later, steam was selected as a working fluid for this analysis. A total of 17.84 lbs/min of water would circulates through the condensor/radiator. The flow rate for the Rankine c cle is approximately 6.0 lbs/min. Thus at the rated cond x tion, the flow of 12 lbs/min would just circulate through the o i l cooler and the radiator only. For this analysis, the steam pressure was assumed to be 1,000 psi. A8 shown later, the pressure can be lowered to 500 psi with only a slight loss in the cycle efficiency. working fluid through various part of the engine is shown below.

Amount of heat to be collected by the

O i l cooler: 2304 Btu/Min To - Recirculatin Water 1652 Btu/Min

652 Btu/Min - Rankine Work 7 ng Fluid Cylinder Head:

Exhaust Manifold:

1130 Btu/Min

357 Btu/Min

Exhaust Stack Boiler: 5250 Btu/Min

Total Heat to Working Fluid: 7389 Btu/Min

Due to the heat energy collected through the engine

19

cooling, the heat transfer at the vapor generator is reduced by approximately 30%. reflected in the size of the vapor generator. Here the oil temperature going into the oil cooler was assumed to be 374'F. capability can be developed , then the reduction can be much higher.

of the process are shown below:

This reduction would be

If a lubricant which has a higher temperature

Physical conditions of the steam at different points

Temperature ("F)

Radiator/Condensor Outlet 236

Water Pump Outlet 236

Oil Cooler Outlet 362

High Press. Feed Pump Outlet 362

Cylinder Head Outlet 544

Boiler Inlet (11.5% Steam) 544

Boiler Outlet (100% Steam) 950

4.2 FLUID SELECTION AND OPTIMIZATION OF

Pressure (PSIA)

30

95

95

1000

1000

1000

1000

RANKINE CYCLE

was Pas

The task of selecting the working fluid for the system granted to Argonne National Laboratory based on their t experience on the subject. Dr. R. Cole selected

water and toluene as two candidates from several possible working fluid candidates. Table 4-1 shows characteristic8 of various organic rankine cycle fluids. Reasons for selecting the above two fluids are described in the attached Argonne report (Appendix 4).

Cycle Analysis/Optimization

Thermodynamic analysis was performed for the following four cases:

1. Steam at 1000 PSI pressure as a working fluid

2. Steam at 500 PSI pressure as a working fluid

3. Toluene at 500 PSI pressure as a working fluid, without regeneration.

2 0

I -

4. Toluene at 500 PSI pressure as a working fluid, with regeneration.

Following assumptions were made in all the analyses:

- Exhaust gas flow rate: 50 lbs/min

- Exhaust gas inlet temp. to evaporator: 1100°F

- Expander and booster pump efficiency: - Expander outlet pressure:

70%

30 psia. A summary of the analyses is shown in Table 4-2. It is clear that the cycle with toluene requires regeneration in order for its efficient to be comparable to the steam cycle. Toluene has considerable amount of energy after expansion in the power cylinder. regenerator would increase complex to the system as well as the manufacturing cost. Therefore, it was decided to select water as a prime candidate for the system. Detailed analyses are depicted in the Appendix 4. thermodynamic (T-S) diagram for the selected cycle is shown in Figure 4-3. Figure 4-4 shows component contributions to the heat input for the integral steam bottoming cycle.

Addition of the

The

4.3 CONCEPTUAL DESIGN OF THE SYSTEM

4.3.1 Heat Exchanger Sizing

Argonne National Lab. was subcontracted for this task and a detailed discussion on this subject is given in the Appendix 4. Here, a brief summary is discussed.

Figure 4-5 shows the size of the evaporator as a function of the working fluid inlet temperature to the evaporator for three levels of the fin spacing. Also shown is the s i z e of the evaporator used by Thermo Electron for their DOE demonstration program. According to a heat exchanger manufacturer whose products include heat exchangers for diesel en ine exhaust gas, it is fins/inch in order to avoid fouling on heat transfer surfaces (Ref. 7). For this analysis, the fin spacin of 6 fins/inch is selected. dimensions are given in Figure 4-6. demonstration unit, the size is reduced approximately by 40%. the base engine.

condensor/radiator.

recommended that the fin spac s ng should be limited to 6 P Details of the tube and f n

Compared to the

However, it is still too large to be attached to

Another heat exchanger required for this system is the Slightly superheated steam from the

21

expander outlet mixes with the engine coolant at or just before the top tank of the radiator. cause the steam to-condense into partially saturated steam since the mass flow rate of the working fluid is 1/2 of the engine coolant. AS shown in the previous section, a total heat rejection of this system is 9041 BTU/Min at a rated 350 HP condition. current diesel enqines with a output power of 350HP is aroulld 9000 Btu/Min, the size of the radiator for this system will be similar to current radiators being used in industry.

The mixing will

Since the heat rejection of

4.3.2 Power Expander

For this system, it is proposed that one of cylinders

The steam reciprocating engine concept has been of the base diesel engine be used as the power recovery device. studied by a few (Ref. 6 and 8), including Foster Miller which performed the study for the NASA/DOE program described in this report earlier. However, when the reciprocator is integrated into the diesel engine, several design issues such as vibration and valve events must be addressed. perform a preliminary feasibility study of the concept, those issues were addressed from that objective in mind. More detailed analyses should be performed in the next phase, if the concept proved to be attractive, or at least worth pursuing further, based on this study.

Since the objective of this study is to

Following are brief comments on those issues:

- Expander Displacement: As described in Table 4-3, a displacement of

approximately 100 cubic inches is required. fit nicely with a cylinder of Cumins L10 engine which was used as a base for this study.

This would

- Intake Valve: It is essential for the efficiency of the expander

that the intake valve opens rapidly to allow the working fluid into the cylinder quickly and have adequate time for expansion work. Therefore, a sliding valve is considered as a intake valve as shown in Figure 4-7. valve event can be more gradual and so can be the same as that of the base diesel engine.

- Vibration/Balance: A detailed vibration analysis is beyond the scope of

this work. configuration of the engine should be made. option is to use the firing order of regular 5-cylinder- engines for the first five cylinders (1-2-4-5-3-1) and the expander cylinder would be positioned at a 180 crank angle

The exhaust

There are several options as to how the One of the

2 2

degree from the NO. 1 cylinder. This would require a counter-balancer at the sixth cylinder to take the first order unbalance out. crank shaft same as the six-cylinder engine with the uneven firing order. Balance due to the inertia forces should be good for this configuration. However, torque on the crank shaft due to gas pressure should be evaluated.

Another option would be to have the

- Lubrication rest of the engine as shown in Figure 4-8. special lubricating oil (Steam engine oil) can be used for the lubrication of components in the sixth cylinder. Those oils are commercially available through Exxon or Mobil (Ref. 5). As a future technology, dry lubricated steam reciprocators should be developed.

The oil sump of the steam cycle is separated from the Therefore, a

4 . 3 . 3 Engine Design/Layout A cross-section of the base diesel en ine is shown in

Figure 4-9 along with a detailed explanat s on of each components. And a schematic, depicting a layout of the engine mounted on a truck, is shown in Figure 4-10.

4 . 4 PERFORMANCE PREDICTION

Based on the thermal analysis described in the sections 4.1 and 4.2, a total engine performance was made and the results are presented in Table 4-4. The best performance is obtained with the turbocharged/aftercooled turbocompound engine plus integrated bottoming cycle and its BSFC can be as low as 0.260 lbs/hp-hr. However, the ,

performance with the non-aftercooled/turbocharged turbocompound engine is also close to the above figure and" shows a good fuel economy.

4.5 MANUFACTURING COST ESTIMATE

Manufacturing cost estimate of the new system was made by comparing the engine against the base turbocharged engine. Appendix 5 of this report. Summary of the cost increase ia shown in Table 4-5. figures to those of the conventional steam Rankine systdm studied under this program, i.e., Foster Hiller s stem.

system. Main contribution of the difference comes from the expander, power train, and the condensor. However, due to its added complexity to the base engine for extracting the additional heat from the cylinder head and exhaust manifold, the cost of vapor generator and engine modification together would cost more for the new system

The detailed analysis of the study is shown in

The table compares the cost

It shows the coat is almost a half of the convent 1 onal

23

than the conventional design. In addition, the new system only generates a 305 HP total output (Turbocharged version) as compared to the baseline engine which generates an output of 317HP. Therefore, when the comparison is made at the 350 hp rating (Here, the 0.7-power law was used to convert the cost at the 350 hp level.), the premium for the new system is increased to $2,400 .

This disadvantage of the lower output can be reduced by increasin the BMEP of the engine. The design point of this part f cular engine was at the BMEP level of 195 psi. As the output level of the base engine increases, the total heat available for the bottoming system increases. Therefore, the output of the expander would increase and the reduction of output power due to the usage of one of the s i x cylinders as a expander would decrease as well. next step if the system is to be pursued further.

This concept should be studied as a

4.6 ECONOMIC EVALUATION

Assumptions regarding to the maintenance cost, operations, etc, made for the evaluation of first three bottoming cycles are also used for this study. Results of the economic analysis is summarized in Figure 4-11. As seen from the figure, the new integrated system would give a much better return on investment for customers compared to any other bottoming systems. However, it still is not as good as the turbocompound engine. When the fuel price exceeds $1.25/gallon range, the new system seems to give a rather attractive investment opportunity for the end users.

24

V. CONCLUSION

Following conclusions can be made from the entire

1. Bottoming Cycles offer good opportunities for large

study described in this report.

fuel'economy gains. However, traditional bottoming cycles are not competitive against turbocompound envines due to its complexity and thus high costs (The initial investment as well as maintenance cost.)

2. A new integrated Rankine/Diesel system was proposed. Based on the preliminary study of the system, it offers the best return on investment among bottoming cycles studied.

3. The new system would give a more than 20% Internal Rate of Return (IRR) at the fuel price of $1.25/gallon0

4 . Therefore, further studies should be made on the new system including:

- further optimization of the concept by studying such areas as,

. optimize the amount of heat recovery for each components. . optimum BMEP of the base engine

concept with an Wltra low" heat rejection engine. - hardware demonstration of the 1-5 steam/diesel

25

120 IYI

40-

2 0 -

Inlet Temperature = 1240OF Cooler Temperature = 115OF Adequate Heater Tube

Engine Weight - Q)g P5J

200 w g

400 .5

Power Available to Cycle -\c f" ' - 1.3 z - 1.1 c Q)

aJ .-

.

- Cycle Power

Brake Power

0.31 A

a 0.30 e

s 0.29

0.28 a

c

k

800 700 800 900 lo00 110 0 V

Stirling Exhau8t Temperature ( O F )

Figure 2-2. RESD V-4 Power Recovery as a Function of Heater Head Temperature (2000 rpm)

28

Regenerator I Cooler

Crankcase

Figure 2-3. SAV-4 Concept

2 9

A

I

Figure 2-4. FPSE Concept

3 0

Baaic Diesel

10 and Rotor Low- -

A I I I I L

.Stirling Compounded

Cycle Power

Figure 2-5. FPSE Optimized a t 2500 rpo and 644 K Heater Tube Temperature

31

r

c -

32

- 6

20

f- stage

Two StageJ d -

e

r Single

d Combined Power

to00 1500 2000 2500 3Ooo Engine Speed (rpm)

Figure 2-7. Comparison of Slngle- and Dual-Stage Power Recovery Based on Mod 11 Geometry

3 3

25

20 i5

E B aa

0 S a x

15

E m

10

5

0 h I I I Y I

0 ' loo0 1500 2Ooo 2500

Heater Head Length

Double

Triple

Quadruple

Engine Speed (rpm)

Figure 2-8. Mod 11 Design Power Recovery as a Function of Heater Head Tube Length

3 4

I

Figure 2-9. Individually Tubed Heater Head Concept

3 5

Figure 2-10. Cold Engine Drive S y S t m

36

I 4'

al m 4 m

C 0 d U cd

U

L, M al

I

37

18.6 in.

Mod II Stirling Engine r / a

Figure 2-12. D i e s e l / S t i r l i n g Engine Integration (End View)

38

2

1300 rpm

1900 rpm

0 50 75 100 Diem1 Load (%)

Figure 2-13. St ir l ing Bottolaing Cycle Patfotnunce Hap with Turbocharged Diesel

39

L PC

HP

R E G

LPT

R EG

Figure 2-14 : S c h e m a t i c of TSA Diesel Eng ine w i t h t h e F o l l o w i n g Nota t i o n : LPC = L o w P r e s s u r e Compressor HPC = High P r e s s u r e Compressor REG = R e g e n e r a t o r ADE = A d i a b a t i c Diesel Enqine STE = S t i r l i n g E n g i n e IlPT = l l iqh Pressure T u r h i n c LPT = Low P r e s s u r e T u r b i n e

4 0

TEMPERATURE

ENTROPY

Figure 2-15 : T-s-Diagram €or TSA Diesel Engine Showing Components and Part Processes. For notation, see following page.

41

LPC

* HPC

REG

DI L

Attachment to Figure 2-15. Notation to T-S Diagram

- Low Pressure Compressor

- High P r e s s u r e Compressor

= l l e a t Regenerator

= Diesel I n l e t Loss

COMPR Compression i n Diesel Engine -

c1-C3 - Combustion i n Diesel Engine

~ 1 - E 3 = Expansion i n Diesel Engine .-

BDL

STIl

IIPT

REG

LPT

ATM

= Blow-Down Lo'sd

= S t i r l i n g Engine I l ea ter

= Iligh P r e s s u r e Turbine

= f i ea t Regenerator

= Low P r e s s u r e Turbine

= Atmospheric P r e s s u r e L i n e

41 -1

O C 2500

2080

1500

1000

500

0

TSA Diesel Engine: Full Load Part Load ---. - 400 11P Adiabatic Diesel Engine: Full Load - --

138SOC

9.2 /MPd --7d /

/ /

/

/’ /

M Pa I 1.4 MPa

2350 O C ----

0.60 MPa

---- 4 6 6 O C

v - ENTROPY 0 0.2 0 . t 0.6 0.8 1 .o 1.2 1.6 k l l k g O C

I . - - - - . - I I - - Figure 2-16 : T-s-Diagram f o r TSA Diesel Engine, Full Load

and Part Load a s well as, for Comparison, 400HP Adiabatic Diesel Engine. ‘FSA Performance below: PerEormance AIM? STI? OV RRAfdL

Power , kW 3 5 H 94 452 Efficiency, P 4 11 40 605 Puel/Ai r Rat io 0.050 - Air Flow, kg/s 0.35 42

-

A D E 110 / STE 4-471

9 3 3 1

9 '0

Figure 2-17. Semi-Integrated Design (In-Line Engines)

4 3

4 > I

I-- 4

t 4 W c v)

c

\

s I 0 -I

W Q

- a I 0

VI 4

I

Lc bo Q - U c n I

-r(

I m

Qo 4 I N

r 4

m

-rl rl A U Y

OI

N -4 I

4 5

C Temperature 4’

2500

2000.

1500.

1000.

500

0

’ t /

I . I I

1/ STIRLING

I -1-

I I I I

I d 1 r i t-I--

1.0 112 1.4 0!2

m Y

-1 m/k8 C

Figure 2-20. T-S Diagram for TSA Diesel Engine with High Temperature Diesel Cycle

4 6

e 0 a 0

4 9

M N 0 U

N I

50

USE OF ONE FLUID.FOR:

- DIESEL ENGINE COOLING - RANKINE CYCLE WORKING FLUID

M n - T 7 Tc; I

- 5 U

21 el 0 P

H I- I I ‘ I

DIESEL ENGm

A SCHEMATIC OF A PROPOSED SYSTEM

Figura 4-1. Integrated Diesel/Ramine Cycle

5 2

EXHAUST OUT -

, TU EXHAUST STACK BOILER

c c

~~l&X$JDER 1 RECIPROCATO

1 CYLlNOER RANKINE

i 0

=! 0

HIGH PRESSURE P WATER euMe

8 si d 5 Y

W

3 FEED-

CONOENSER / RADIATOR

Figure 4-2. Integrated Steam Based Rankine Bottoming Cycle Schematic

5 3

,

LMTRWV. I, #IWLW =

Figure 4 - 3 . T-S Diagram for the Steam Rankine System

5 4

A-B INTEGRAL WITH ENGINE

C-0 INTEGRAL WITH ENGINE D-E VEHICLE MOUNTED IN AIRSTREAM'

e-c EXHAUST STACK BOILER

Figure 4-4. Steam Cycle Showing Integral Bottoming Cycle Component Contributions

5 5

10.0

8 . 0

6.0

4.0

2.0.

' 0 i

fins /inch

fins / inc h

fins /inch

I I 1 I 1 200 300 4 00 500 600

Warklng Fluid Ialet Tcrpcrature to the Ewaporatoc, .P

Figure 4-5. Evaporator Size for the Integrated WM bottombg Cycle for Truck M c s d -ne - Steam e lo00 PSI

5 6

I L

Tube

Outside Diameter = 1/2" o r 5 / 8 "

Wall thickness = 0.04" t o 0.05"

Material = carbon s t e e l

p i t c h p = 3 / 4 "

-

Fins

Thickness = 0.015" to 0.020" -

Material = low carbon steel

No. of f in s l inch = 6

Multiple tube and f i n compact

Heat exchange design is recommended.

Brazed tube to f i n j o i n t should be used

Overall dimensions t o s u i t packaging on the engine

Tota l Fin:. surface area should match the f igu re

chosen from the graphs.

Figure 4-6. Evaporator Tube and Fin Details

57

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F i g u r e 4 - 8 . Crank/Sump Layout f o r t h e Steam C y l i n d e r 5 9

I - i. ~

CF

F

I) -

Figure 4-9. Cross Section of Diesel Reciprocator f o r Integral Bottoming Cycle Englne (See fo'll-owinq page for corresponding design features)

6 0

Attachment to Figure 4 - 9 .

DESCRIPTION OF DESIGN FEATURES FOR INTEGRAL BOTTOMING CYLCLE ENGINE (See sheet F for corresponding illustration)

PAGE 1 OF 2

A. Isolated Unit Pump Assembly. Isolated to reduce heat rejection from the head to fuel and for service. External to valvecover. 115 degree f fuel temp to injection nozzle. PT or ECI controled.

8 . Fuel Nozzle. Isolated from head and valvetrain for reduced heat rejection. Cooled only by fuel used for combustion. Stanadyne type slim tip nozzle end features reduced tip area exposed to combustion chamber. Externally serviced. Surronded by an air gap, exposed to ambient . c. Overhead Cam Valvetrain Assembly.

c1. Pivoting valve and fuel injection rocker arms. c2. Overhead cam retained by split bearing inserts and

bolt on caps. C3. Valve gear pedestal. Bolts to head as an assembly. C4. Cam cover C 5 . Valve cover, provides access to mechanism

adjustments, locates’fuel nozzle.

D. Thick insulativ blanket. Surrounds all areas used for heat recovery. convection.

Reduces heat loss by conduction,

E. Engine Breathing System. El. Isolated inlet charge plenum. Provides reduced

heat rejection to charge air. minimized.

E2. Exhaust port/manifold retains pulse geometry and uses cast in high pressure steam/water tubes.

Inlet head port area

F. Iron Cylinder Head. Cooled only b boiling condensate

incident valvetrain oil draining.

G. High Pressure Condensate Inlets. For increased strenght, the cast in steeltu6es feature flared ends so that high prepsure nipples may,be threaded directly into the steel tubes.

in steel tube passages cast directly x nto head, and by

G1. Block Inlet. ( 1 per c linder ) G2. Head Inlet ’( 1 per cy1 1 nder )

H. Engine Block. Simplified, incorporating linerless design, cooled by condensate boiling in cast in tubes in the top ring reversal area. Additional cooling provided by an underside piston/liner oil spray.

61

OF 2

, J . z i r c o l ~ i a s p r a y coated d u c t i l e i r o n o r CLAS cast steel p i s t o n . cool ing provided by forced o i l m i s t on t h e unders ide . Convent ia l 3 r i n g des ign .

K. D1ow;ny and c y l i n d e r to c l i n d e r b r e a t h i n g chambers.

t o the scavenge pump.

L. High temperature l u b e oil system.

Also a c o l e c t i o n p o i n t for h x gh t empera tu re o i l t o r e t u r n

L1. Valve t r a in and head o i l r e t u r n tube . ~ 2 . Pis ton/Liner l ube /coo l ina t u b e and connec t ion .

M. I n s u l a t j v e Gasket System. U t i l i z e s a steel and PSZ sandwich to i s o l a t e h igh and l o w t empera tu re components.

M I . Injector t i p asket . PSZ cone seals combustion

M2. I n l e t plenum gaske t . PSZ or e q u i v i l e n t t h e r m a l l y

premure and P solates f u e l cooled n o z z l e from cy1 i n d e r head . isolates t h e c o o l i n l e t plenum and charge a i r (110 deg F) from t h e c y l i n d e r head.

N. Ded/Crankcase sump assembly. The bed assembly s u p p o r t s the c r a n k s h a f t i n t h e normal manor and forms t h e top half of the sump.

P. Main Cap/Ded/Block s t u d s . For s i m p l i c i t y and improved a l ignment , a s ing le set of s t u d s is used t o t i e t h e main bea r ing cap , bed, and b lock together t o form a s i n g l e r i g i d u n i t .

6 2

I

L-- - - ELECTRONIC I 1 CONTROL MOWLE (CAB MOUNTED1 /

I I

Figure 4-10. Schematic of Vehicle Mounted Equipment for Integrated

63 Bottoming Cycle Engine (Overhead View)

J t- i? W G a A Q 7 U W

0 a0

0 CD

0 cu 0 0 cu I

0 I- c3 A k UJ

\

0 I- c3 \

8 0 I- t 6 W

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3 \ n n 0 I-

6 4

TABLE 2-1. MAINTENANCE SCHEDULE (HOURS)

~~

Configuration

Descri p t ion nod I1 SAV-4 FPSE

Bearing Relubrication - 5,000 - 5,000 - - Piston Seal

Piston Rings

Shaft Face Seals

5,000 5,000 - 5,000 2,000 -

- Uain Bearing Replacement 5,000 - Change Oil

Flush Coolant

OK/Hydraulic Filter

5,000 - - 500 500 5 00

10,000 - 10,000

- Controls (External Check Valve) 10,000 10,000

H2 Supply 5,000 5,000 -

6 5

TABLE 2-2. MAINTENANCE SCHEDULE (COST)

Hod I1 SAV-4 FPSE

9

Bearin8 Relubrication .Material Labor*

Piston Seal Material Labor*

Piston Rings Haterid Labor

Shaft Pace S e d Ha t er ia 1 Labor*

Main Bearing Replacement Haterial Labor*

Oil Change Hater ial Labor

Pluah Coolant Ha t et i 1 1 Labor

Oil/Bydrogan Filter Hat et i a1 Labor

Control a Hater i a1 Labor

HZ Supply Material Labor

Tear Down Engine Labor

$3 Included

$2O/Set of 4 Included

$28/Set of 4 Inc luded Inc ludcd

$2O/Set of 4

$168/Set of 7 Inc 1 uded

$ 5 / 4 Quarts $26

$6/Callon $6/Callon $6 f Gal Lon $52 $52 $52

$8 $9

$ 5 $17

$8 $9

$5 $17

$560 (16 hr) $420 (11 hr)

*Hain Labor Cost is Included in Engine Tear-Down Cost Production Rate of 10,000 Units/Year Based on $35/hr Labor Rate 6 6

TABLE 2-3. ljAINTENANCE SUMMARY

Evaluated Over the Life (1400 Hours) of t h e Engine

MOD 11 SAV-4 FPSE This inc ludes two complete changer of

P i s ton Seals

This inc ludes two Includes (3) changes complete changes o f : of coolanc

and one change of 0 BRC Lubricat ion Hydraulic Filter

0 Pis ton Rings

a Shaf t Seals

0 Piston Rings

0 Engine tear down

0 Main Brgr a Oil

plus one change of

con t ro l va lves

0 Hydrog8a Plus (7) changes o f : 0 Engine tear down plum

0 O i l f i l t e r

0 Coat ro l valve

0 Shaft Seals 0 Hydrogen one change of:

p l u s (3 ) coolant f l u r h e r

MAT'L - $ 504

LABOR - $1,389

TOTAL - $1,893

MAT'L - $ 884 MAT'L - $ 30

LABOR - $1,8S9 LABOR - $ 174

TOTAL - $ 2 , 7 4 3 TOTAL - $ 204

67

TABLE 2-4 COMPARISON OF THE THREE ENGINE CONCEPTS

(BY MTI)

MOD-I1 SAV-4 FPSE

*1 FUEL SAVINGS ( % ) 10 11 a

A- "&

MANUFACTURING COST $2,580 $1,940 $7,000

MAINTENANCE COST $1,893 $2,743 $204 *3

WEIGHT (LBS) 539 781 620

TECHNICAL RISK L O W HIGH VERY HIGH

DIFFICULT VERY DIFF PACKAGABILITY BEST. '

OTHER FACTORS SMOOTH TORQ ROUGH TORQ HIGHLY COMPLEX

OVERALL ~

BEST CHOICE

*1: BASED ON TURBOCHARGED CORE ENGINE AND OTHER PARAMETERS ARE STIRLING EXHAUST GAS TEMPERATURE: 700'F WORKING FLUID: HYDROGEN. STIRLING ENGINE SPEED: 1,000 RPM COOLER TEMPERATURE: 115'F

*2: ROUGH ESTIMATES BY MTI (10,000 UNITS PRODUCTION/YEAR)

*3: FOR A 7-YEAR LIFE TIME

6 8

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TADLE 2-6 : IIEAT DALANCE FOR TSA DIESEL E N G I N E :

C o n d u c t i o n Losses (103: of F u e l Meat I n p u t ) 74.7

Diesel. m g i n e S h a f t Power ( E x c l u s i v e of F r i c t i o n ) 357.5

Diesel E n g i n e F r i c t i o n L o s s e s 1 1 . 1

Diesel E x h a u s t Heat to S t i r l i n g E n g i n e 235.3

S y s t e m E x h a u s t Loss (From 210 to 29.5 D e g r e e s C ) 67.2

Total llcat O u t of Diesel Enq . ine

C a l c u l a t e d I n p u t of F u e l Heat 6 0 2 . 7

Added Heat I n p u t to C o m p e n s a t e f o r t h e Use of

E x t r a p o l a t e d c - V a l u e s 144.7 P

T o t a l Heat F low Into Diesel E n g i n e

T o t a l Heat Plow I n t o S t i r l i n q E n g i n e

S t i r l i n q Encj inc ShaCt Power ( I l f f i c i c n c n / = 4 0 % )

235.3

94.1

7 0

TABLE 3- 1. ECONOAIC/OPERATIONAL ASSUMPTIONS

- Tax A Corporate Tax Rate ........ 3 4 1 ,

. Investment Tax Credit ........ 0%

*1 - Equipment Price/Salvage Value/Life . Annual Production Rate ........ 10,000 units . Selling Price/Mfg Cost . e . . . . . . 2.0 - . Future Cost Reduction ........ 30%

(Learning Curve Effect) I .

. Salvage Value ( b of Original) ...... 20%

. Hardware Useful Life ........ 7 years - Fuel Economy

. Anuual Truck Mileage ........ 100,000 miles

. Diesel oil Price ........ $1.00/Gal *2

. Base Engine (TCPD/AC) Truck Fuel Mileage .o...... 8.0 MPG *2

*1: Assumed a 10% penetration on a market of 100,000 class-8 trucks

*2: Sensitivity Analysis was made around this base case

7 2

(VMS STUDY)

ORGANIC RANKINE

STEAM RANKINE

STI RL 1NG4

BRAYTON

RATED HILLYl ELBz2 Mu3 13 I 9% I3 e 5% 13,7%

13.2% 13.3%

(15.0%)

(14.1%) 13.6%

(10.0%) 9.4% 8.9% 9.1%

10.6% 9.5% 10.3% (11.1%)

RENO----SACRAMENTO

I NDIANAPOL IS----CH ICAGO

COLUMBUS----LOU ISVILLE----CINCINNATI ----COLUMBUS

' AUTOMOTIVE STIRLING ENGINE IS USED (FOR SIZE 8 COST), LARGER ONE WOULD GIVE 12.9% IMPROVEMENT AT RATED CONDITION.

73

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7 4

TABLE 3-4 .

MANUFACTURING COST OF BOTTOMING CYCLE SYSTEMS

MAJOR COMPONENTS RANKINE STIRLING ORGANIC STEAM

1. VAPOR GENERATOR OR HEATER HEAD

$1,200 $800 $534

2 . EXPANDER/HOUSING $863 $1,336 $939

3 . CONDENSOR/REGENERATOR $690 $650 $734 (INCL. FAN & OIL COOLER)

4. POWER TRAIN/CLUTCH $538 $470 $470

5 . WORKING FLUID SYSTEM $1,030 $472 $146 (PLUMBING,PUMPS,ETC.)

6 . CONTROL SYSTEM $305 $305 $254

7. ASSEMBLY/ PRE-SHIPMENT TEST

$312 $166 $203

TOTAL $4 , 9313 $4 , 199 $3 , 258

OPTIMISTIC ESTIMATE ($3 ,457) ($2 ,939) ($2 ,281)

7 5

TABLE 3-5. ENGINE SYSTEM PRICE TO CUSTOMERS

Turbocharge W j

Turbocompound (TCPD)

TCPD/Aft . Cool (TCPD/A)

(Q 350 HP level)

Rankine Base Steam Organic

$15,005 $23,192 $24,670

$15,439 $23,692 $25,170

$16,121 $24,826 $26,304

$16,134 $25,326 $26,804

( $ 2 u , i 3 i j * ($21,469)

($20,934) ($21,969)

($22,068) ($23,103)

($22,568) ($23,603)

Stirling

$21,210 ($19,047)

$21,710

$22,844

$23,344

($19,547)

($20,681)

($21,181)

* Numbers in parenthesis indicate 880ptimistic8t value. - Price is based on the 0 .7 power law (1)

Price (350 HP) = Base Enqine Price * 0 . 7 (350.0/Combined System Output)

- The price includes the installation costs at OEM. Installation cost data are provided - by Navistar

7 6

TABLE 3 - 6 . A SUMMARY OF MANUFACTURING/MAINTENANCE COSTS

STIRLING RANKINE

ORGANIC STEAM

MANUFACTURING COST $3 , 258 $4 , 938 $4 , 199 ($1, 789) *1 ($4 , 189) *2 ($3 , 035) *2

MAINTENANCE COST $737 $1,034 $850

($431) (1,100)*2 ($580) *2

*1: NUMBERS IN PARENTHESIS ARE ESTIMATES BY SUBCONTRACTORS FOR

EACH BOTTOMING SYSTEM, I.E., MTI, TECO, AND FOSTER-MILLER

*2: FIGURES ARE BASED ON A 1983 $ VALUE.

77

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TABLE 4 - 2 . RESULTS OF-THE CYCLE ANALYSIS FOR THE DIFFERENT CASES

Work1 ng Flufd

Operating Cycle Pressure Efficiency

1. Superheated Steam 1000 19.8

2. Superheated Steam 500 19.. 5

3 . Superheated Toluene wlth Regeneration 500 18.0

4. Superheated Toluene without Regeneration 500 13.6

Expander Flow Rate Power lbm/min H e P.

5.98 34.9

6.10 34.4

32.40 32.1

24.30 24.1

8 0

TABLE 4-3. CALCULATION OF EXPANDER DISPLACEMENT

Specific volume of steam at exapnder outlet - 16.44 C. - f t lb

Steam flow rate = 6 lb/min.

= 6 x 16.44 c.ft 98.64 c . f t lain ain

If the engine is reated at 1900 rpe, Steam flow rate = 98.64 c.ft 0.0519 toft

1900 rev rev.

One exh. stroke/rev. displacement = 0.0519 c . f t stroke

= 89.68 c o i n 90 c . i n Vol. 0.9, Dieplacement = 100 c o i n

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82

T A X E 4.5. MANUFACTURING COST OF STEAM RANKINE BOTTOMING CYCLE SYSTEMS

MAJOR COMPONENT CONVENTIONAL INTEGRATED (@350HP)

1. VAPOR GENERATOR ENGINE MODIFICATION

$800 $700 $772 $353 $389

2. EXPANDER/HOUSING $1,336 $326 $360

3. CONDENSOR $650 $307 $339 (INCL. FAN $ OIL COOLER)

4. POWER TRAIN/CLUTCH $470 $0 $0

5. WORKING FLUID SYSTEM $472 $315 $347 ( PLUMBING, PUMPS, ETC. )

6. CONTROL SYSTEM $305 $125 $138

7. ASSEMBLY/ PRESHIPMENT TEST

$166 $50 $55

$4,199 $2,176 $2 ,400

8 3

APPENDIX 1

M.T.I.

COST STUDY OF STIRLING

BOTTOMING CYCLE SYSTEM

8 5

:HEATER i YEAD

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APPENDIX 3

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APPENDIX 4

INTERGRATED BOTTOMING CYCLE FOR

TRUCK DIESEL ENGINES

125

ARGONNE NATXONAL LABORATORY 9700 SOUTH CASS AVENUE

ARGONNE, I L L I N O I S 60439

INTEGRATED BOlTOLIINC CI[W FOP TRLTCK DIESEL ENGINES

R.R. Sekar and R . L . Cole Energy and Environmental Systems D i v i s i o n

February 1 9 8 7

Work sponsored by

Cummfns Engine Company Columbus, Indiana

under Work for Others Contract No. 3 1 - 1 0 9 - E N C 3 8 - 8 5 4 0 2

126

CONTENTS

L i s t of Figures

L i s t of Tables

Abstract

1 . Intrpduction

2. Description of the Concept

3. Se l ec t ion of Working Fluid

4. Cycle Analysis

5 . Heat Exchanger Considerations

6. Other Design Considerations

7 . Conclusions

8. References

127

LIST OF FIGURES

1.

20

3.

4.

5 .

60

7.

8 .

9.

10.

Schematfc Diagram of t he Organic Rankine Bot toming Cycle Demons t r a t i o n i n a DOE Test Vehicle

In t eg ra t ed Rankine Bottoming Cycle (IRBC) Concept

,T-S Diagram of IRBC - Steam a t 500 PSI as Working F lu id

T-S Diagram of IRBC - Toluene a t 500 PSI as Working Fluid

T-S Diagram of IRBC - Toluene a t 500 PSI with Regenerat ion

T-S Diagram of IRBC - Toluene a t 500 PSI, no Regeneration

Evaporator S ize f o r t he I R E f o r Truck Diesel Engine - Steam a t 1000 PSI

Evaporator Size f o r t h e IRBC - steam a t 500 PSI

Evaporator S ize f o r t he IRBC - Toluene a t 500 PSI

Evaporator Tube and Fin Details

12 8

1.

2.

3.

4.

5 .

6 .

7.

a.

9.

LIST OF TABLES

Fluid Characteristics

Thennoel Stability - Rules of Thumb Definition of Toxicity

Definitions of Fire and Explosion Hazard

Additional Fire Hazard Data

Cycle Analysis Calculation for Steam at 1000 PSI

Results of Cycle Analysis for other cases

Heat Exchanger Calculation Examples

Calculation of Exapnder Displacement

129

1

Integrated B o t t o d a g Cycle for Truck Diesel Engfnee

R a j Sekar and Roger L. Cole

This s tudy w a s undertaken t o assess t h e f e a s i b i l i t y of Incorpora t ing a RanKlne Bottoming Cycie as park or' a iruck i y p a d i e s e l engine. The Organic Rankine Bottoming Cycle (ORBC) t h a t was previous ly demonstrated by the Department of Energy (DOE) i n a heavy duty long haul t ruck showed about 12% improvement i n f u e l consumption. However, t h a t system was considered t o be t oo complex and c o s t l y t o be commercialized. The i n t e g r a t e d system descr ibed he re is an at tempt t o s i m p l i f y and reduce t h e c o s t of t h e ORBC system. The main f e a t u r e s of t he i n t e g r a t e d system are: 1. One c y l i n d e r of a s ix c y l i n d e r t ruck d i e s e l engine w i l l be used fo r power recovery, r a t h e r than t h e t u r b i n e and reduct ion gears employed i n t h e prev ious ORBC system. 2. Same f l u i d w i l l be used f o r engine cool ing and as working f l u i d i n t h e bottoming cycle. 3. The r a d i a t o r used t o coo l t h e engine coolan t will serve as t h e condenser f o r t h e bottoming cycle as w e l l . Toluene and steam were considered i n t h i s assessment and i t was concluded t h a t steam wil l be more p r a c t i c a l working f l u i d . Steam at 1000 p s i , p a r t i a l l y vaporized t o about 33% s a t u r a t i o n i n the c y l i n d e r head and superheated i n t h e evapora to r , is t h e recommended working f l u i d . The hea t exchanger s izes are smaller than t h e p rev ious ly demonstrated ORBC system but s t i l l may pose a cha l lenge i n under t h e hood i n s t a l l a t i o n of a t ruck. Design and l ayou t drawings and c o s t comparisons are beyond t h e scope of t h i s e f f o r t by ANL and are being done s e p a r a t e l y by t h e sponsors. Overa l l t h e concept appears t o be f e a s i b l e .

Rankine cycle has been t h e mainstay of i n d u s t r i a l and u t i l i t y power

g e n e r a t i o n for over a century. I n t h e s e a p p l i c a t i o n s f u e l is d i r e c t l y burned

i n b o i l e r s t o gene ra t e steam, which is then used f o r d r i v i n g a prime mover t o

g e n e r a t e power. Simultaneously d i e s e l engines were developed t o be a highly

r e l i a b l e prime mover for t r a n s p o r t a t i o n and s t a t i o n a r y a p p l i c a t i o n s . It Fs

g e n e r a l l y w e l l known t h a t both O t t o cycle and Diesel engines have thermal

e f f i c i e n c i e s i n the 25%-40% range. This means t h e remaining f u e l input energy

of about 60% or more is l o s t t o t h e ambient through t h e coo lan t and t h e

exhaust gases. Reacting t o the petroleum price e s c a l a t i o n s and shor tages of

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2

t h e 1970's, governement and industry s t a r t e d r e sea rch work on u t i l i z i n g the

wasted exhaust energy from engines t o gene ra t e u s e f u l power. The U.S.

Department of Energy funded t h e development and demonstrat ion of t h e

turbocompound system (Ref. 1,2,3) and Organic Rankine Bottoming Cycle (ORBC,

Ref. 4 , 5 , 6 ) f o r long haul heavy duty d i e s e l t r u c k a p p l i c a t i o n s , The e f f i c i e n t

recovery of waste hea t becomes even more cr i t ical f o r a d i a b a t i c d i e s e l engine

concept , which has been g e t t i n g considerable a t t e n t i o n i n r ecen t years. The

complexity and c o s t of implementing t h e ORBC system i n a d i e s e l t ruck have

been t h e main reason f o r t h e indus t ry ' s r e luc t ance t o commercialize t h e

concept. A comparative eva lua t ion of t h e waste hea t recovery systems is

r epor t ed f n Ref. 7. It is clear t h a t un le s s s i g n i f i c a n t cost reduct ion - and

s i m p l i f i c a t i o n s are demonstrated, ORBC w i l l not be a t t r a c t i v e t o ' t h e

indus t ry . Cumins Engine Company and Argonne n a t i o n a l Laboratory (ANL)

undertook t h i s f e a s i b i l i t y s tudy t o make t h e Rankine Bottoming Cycle (RBC)

more pract ical and a t t r a c t i v e f o r commercialization,

2. DESCRIPTION Op TEE COWCePT

The bas i c ORBC system is shown schemat ica l ly i n f i g u r e 1. The working

f l u i d forms a s e p a r a t e loop with its own evapora tor and condenser. The engfne

exhaust gas is the source of heat and a power t u r b i n e extracts work from t h e

working f l u i d and, through reduct ion gears, f eeds t h e power t o t h e crank-

s h a f t . The f e a t u r e s included i n the i n t e g r a t e d s y s t e m are:

a ) The power recovery turbine and r educ t ion gea r s are e l imina ted and

one oE t he e x l s t i n g power c y l i n d e r s is used f o r the bottomfng

cycle .

131

3

b) The working f l u i d f o r t h e bottoming c y c l e is t h e same as the engine

cooling f h i d . These two f e a t u r e s have system s i m p l i f i c a t i o n and c o s t r educ t ion

p o t e n t i a l s . The engine coolan t is pre-heated i n the head and t h e r e f o r e has

t h e p o t e n t i a l t o reduce t h e s i z e of t h e evaporator . The schematic diagram of

L L - * - L - - - - & - J ---I-*-- L-ccerl-n nw-ls 4 - a k n e . m L h i c A i i L e E L a L c u L a t t I c A t L s u v c ~ v u * . - - ~ rj--.- +Y in fig:== 2. S%EC= 3 I .Z=~D,

percentage of t ruck d i e s e l s have a f t e r c o o l e r s , a i r - to-air hea t exchangers

would be used for t h a t purpose. This approach w i l l a l low t h e ,coolant t o boil

in t h e engine without adverse ly a f f e c t i n g i n t a k e a i r temperatures. O i l coo l ing

might be done with engine coolan t without adverse e f f e c t s . A por t ion of t h e

coo lan t is d ive r t ed to t h e i n l e t of t h e booster pump where t h e coo lan t is

pres su r i zed to 1000 p s i before i t e n t e r s t he engine cy l inde r head. While

coo l ing t h e head, t he coolan t is vaporized and steam a t an es t imated q u a l i t y

of 33% would come ou t of t he head. This working f l u i d is then routed through

a waste hea t recovery hea t exchanger where superheated steam a t 1000 p s i is

produced from the energy in t h e engine exhaust gases. This superheated steam

is then expanded in one c y l i n d e r of the engine. The power generated f n t h i s

c y l i n d e t i s desfgned t o be 1/6 t h of t he r a t e d engine power of a s ix c y l i n d e r

t r u c k engine. The exhaust steam is then routed t o t h e t ruck r a d i a t o r , where

it mixes w i t h the rest of the coolant from the engine and condensed. In o r d e r

t o improve t h e thermal e f f i c i e n c y , sometimes a regenera tor is a l s o included in

t h e system. However, i n t h i s a p p l i c a t i o n i n s t a l l a t i o n space is l imi t ed .

Hence use of a r egene ra t ive hea t exchanger i n t h e cyc le is not advisable .

132

4

3. SELE(ZI0N OF IUIBKING FLUID

Severa l o rganic f l u i d s have been considered f o r use in t h e bottoming

cyc le . Most common among them are Toluene, F luo r ino l 85 and RC-1 (60% penta-

fluoro-benzene, 40% hexa-fluoro-benzene). The DOE demonstration t ruck used

Fluori-nok 8 5 . Obviously water has t o be cons idered as a candidate. In

a d d i t i o n t o t h e shape of t h e temperature - entropy diagram, t o x i c i t y ,

deg rada t ion temperature , products of decomposition, f i r e hazard and o t h e r

phys i ca l p r o p e r t i e s of t he f l u i d must be considered in t h e s e l e c t i o n of t h e

working f lu id . S ince a v a i l a b i l i t y and c o s t are always important c o m e r c i a 1

concerns, t hese c h a r a c t e r i s t i c s should a l s o be kept i n mind. Since t h e

working f l u i d is t h e same as engine coo lan t , t h e s e l e c t e d f l u i d should a l s o

have good s p e c i f i c hea t and o the r p r o p e r t i e s r equ i r ed of a coolant . From

l i t e r a t u r e sea rch and previous work done a t ANL, t he fol lowing t a b l e s (Tables

1-5) were prepared f o r a few candidate f l u i d s .

Table 1 con ta ins c h a r a c t e r i s t i c s of va r ious o rgan ic rankine cyc le

f l u i d s . The s i g n i f i c a n c e of the var ious columns is as follows:

Average molecular weight: A h igh molecular weight i m p l i e s a dense

vapor in . t h e condenser and a smal le r , less c o s t l y condenser. A high p res su re

a t 220°F a l s o impl ies a dense vapor in t h e condenser. I f t he condensing vapor

were a p e r f e c t gas , i ts s p e c i f i c volume would be:

1 m. w.

v =-

where m.w. is t he molecular weight, R is t he u n i v e r s a l gas cons t an t , T is t h e

a b s o l u t e temperature , and p is t h e abso lu te pressure . Although the condensing

vapor is not a p e r f e c t gas , the per fec t gas assumption g ives a rough estimate

of s p e c i f i c volume t h a t is adequate f o r ranking t h e va r ious f l u i d s .

5

Xaxlmm use temperature: A high maximum use temperature i m p l i e s a high

t h e o r e t i c a l cyc le e f f i c i e n c y and maximum e x t r a c t i o n of usable energy from the

waste heat. Only the va lues lor Fluor ino l 8 5 , to luene and water are w e l l

known;, values f o r t h e o t h e r f l u i d s l i s t e d should be considered o p t i m i s t i c

because ,they are based on s t a t i c capsule tests. Table 2 g ives some r u l e s of

thumb f o r e s t ima t ing t h e decomposition temperatures. Maximum use temperatures

will be 100-300°F less than t h e temperatures g iven i n Table 2.

Given a n llOO°F exhaust temperature and a 220-25O0F condensing

temperature , t h e bes t f l u i d s can be expected t o have about a 20% t h e o r e t i c a l

c y c l e e f f i c i ency . Therefore , f l u i d c h a r a c t e r i s t i c s o t h e r than t h e o r e t i c a l

cyc le e f f i c i e n c y are l i k e l y t o have a major i n f l u e n c e on the choice of f l u l d .

Plow or Freezing P o i n t p l aces a lower l i m i t on eys t en o p e r a t i o n

a l though an automatic drain-down system could be designed f o r a water system.

I-factor is formally def ined as:

where D r e f e r s t o t h e dewpoint line. F l u i d s wi th I greater than about 1.0-1.3

may not opera te s a t i s f a c t o r i l y w i t h t u r b i n e expanders , but may be accep tab le

w i t h posi t ive-displacement expanders (Le . piston-, vane-, o r s c r e r t y p e s ) . A

r egene ra t ive hea t exchanger will be r equ i r ed f o r small I - fac tor f l u i d s , bu t

f o r I - f ac to r s g r e a t e r than 1.0, t h e r e g e n e r a t i v e hea t exchanger may not always

be required.

Pressure at 220% I n a d d i t i o n t o g i v i n g information on the condenser

volume, (see a l s o average molecular weight ) , t h e p re s su re a t 220°F g ives an

i n d i c a t i o n of whether a i r could l eak i n t o the s y s t e m and ox id ize t h e f l u i d o r

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6

whether t h e f l u i d could leak out through s h a f t seals, p i s t o n r ings , va lve

packings, and so for th .

T o x i c i t y g ives an ind ica t ion of hazard t o personnel. Table 3 (from I.

Sax, Dangerous P rope r t i e s of I n d u s t r i a l Materials) d e f i n e s the var ious l e v e l s

of t o x i c i t y as used i n Table 1. For comparison purposes, gaso l ine is r a t e d a

moderate-to-high hazard v i a inha la t ion .

F i r e a d Bxplouton Bazardr are f u r t h e r i n d i c a t i o n s of hazards t o

personnel . Table 5 (from I. Sax, Dangerous P r o p e r t i e s of I n d u s t r i a l

Materials) d e f i n e s these hazards. For comparison, gaso l ine is a dangerous

f i r e hazard and a moderate explosion hazard. NO. 2 d i e s e l f u e l is a dangerous

f i r e hazard. Addi t iona l f i r e hazard d a t a is given i n Table 5 .

TOXIC Dccorposi t ion Product8 & Toxic P a r t f r l Oxidation Product.: These

products are l i s t e d where they are known even i f they are produced i n very low

concent ra t ion .

From an a n a l y s i s of known prope r t i e s d i scussed above, water and Toluene

were chosen as t he two p r a c t i c a l f l u i d s f o r t h i s app l i ca t ion . Water is a

benign l i q u i d t h a t is eas i ly a v a i l a b l e and has been widely used both as

coo lan t and Rankine Cycle working f l u i d . Hence its acceptance w i l l be easy.

Besides steam r e c i p r o c a t o r s are well understood i n practice. However the re is

one important concern with water: it f r e e z e s a t 32 F. Trucks commonly use 50%

e t h y l e n e Glycol - Water Hfxture. This mixture as bottoming cyc le working

f h i d would create problems with high temperature ox ida t ion and decomposition

products . This i s s u e has t o be recognized and solved du r ing experimental

phase of t h e pro jec t . For t h i s f e a s i b i l i t y s tudy , steam and pure water

p r o p e r t i e s were used. Among t h e organic compounds considered to luene appears

t o be the most s u i t a b l e . A v a i l a b i l i t y and c o s t of to luene are reasonable ,

p r i m a r i l y due t o the i n d u s t r i a l research on t h i s f l u i d over many years. Hence

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7

water is recommended as the primary f l u i d and to luene is t he second choice f o r

t h i s app l i ca t ion .

6. CYCLE ANALYSIS

Thermodynamic cyc le a n a l y s i s was performed f o r t he fol lowing cases:

1. Steam at 1000 p s i p re s su re as working f l u i d . Four

l eve l s of feed water preheat were considered.

2. Steam a t 500 p s i p re s su re as working f l u i d . Four

l e v e l s of feed water preheat were considered.

3. Toluene a t 500 psi as working f l u i d , without

regenerat ion.

4. Toluene a t 500 p s i as working f l u i d , wi th regenerat ion.

The followtng assumptions were made in a l l t h e ana lyses :

Exhaust gas flow rate---------- 50 l b s / m i n

Exhaust gas Lnle t temp t o evapora tor 1100 F

Expander e f f i c i e n c y 70%

Booster pump e f f i c i e n c y 70%

Expander o u t l e t p re s su re 30 p s i

The various cyc le s analyzed are presented i n f i g u r e s 3 - 6 . The cycle

c a l c u l a t i o n s a re descr ibed i n f u l l d e t a i l f o r t h e case of steam a t 1000 p s i .

136

All o t h e r cases followed similar logic.

S t a t e point 1 r e f e r s t o water a t t he i n l e t t o t he boos te r

pump, 30 p s i p ressure , l i qu id s t a t e

Skate poin t 2 r e f e r s to water at the o u t l e t of t he boos te r

pump, LO00 p s i pressure, l i qu id state

S t a t e poin t 3 corresponds t o t h e pure l i q u i d state of 1000

p s i steam. The process represented by 2 - 3 is t h e feed

water prehea t ing phase, which is accompllshed to varying

degrees in cool ing t h e cy l inde r head. Cummins es t imated

t h a t t he maximum preheat ing t h a t could be accomplished

wi th in t h e engfne is t o generate 33% q u a l i t y steam, which

then can be superheated in an evaporator . The main impact

o f prehea t ing t h e working f l u i d in t h e engine is t o reduce

the s i z e of the evaporator .

S t a t e po in t 4 corresponds t o t h e s a t u r a t e d steam state of

the working f l u i d . The working f l u i d is brought t o t h i s

s ta te by a combination of p rehea t ing in t h e engine and

evapora t ion in the bof ler.

S t a t e po in t 5 is t h e superheated s ta te of t h e working

f Luid. Process 4-5 is t h e superhea t ing process

accomplished i n t h e evaporator.

9

Process 5-6 is, t h e expansion of t h e working f l u i d i n one of t h e power

c y l i n d e r s of t h e engine. It is i n t h i s process t h a t exhaust energy is

recovered as u s e f u l power. S t a t e po in t 6 is s e l e c t e d t o be on t h e 30 p s i

p ressure l i n e and t o conform t o our assumption of 70% i s e n t r o p i c e f f i c i e n c y .

Process 6-7 is t h e condensat ion of t h e superheated steam from the power

c y l i n d e r exhaust t o the state of s a t u r a t e d steam. Process 7-1 is t h e

cont inued condensation t o pure l i q u i d state. The e n t i r e condensat ion process

is t o be accomplished i n the r a d i a t o r .

The d i e s e l engine exhaust gas e n t e r s t h e evapora tor a t 1100 F and

t r a n s f e r s the energy t o the working f l u i d as shown by t h e l i n e marked "exhaust

gas". The s lope of t h i s l i n e is chosen t o provide reasonable temperature

d i f f e r e n t i a l i n t he evapora tor , e s p e c i a l l y a t the "pinch point" , which is t h e

s ta te poin t 3.

The cycle a n a l y s i s f o r t he case of steam a t 1000 p s i is shown i n d e t a i l

in Table 6. Resu l t s of similar a n a l y s i s f o r t h e o t h e r cases are shown i n

Table 7. It i s important t o note t h e d i f f e r e n c e i n t h e shape of t h e

s a t u r a t i o n curves f o r steam and toluene ( f i g u r e s 1 and 3). It is due t o t h e

shape of the curve t h a t to luene r e q u i r e s a r egene ra to r t o avoid e f f i c i e n c y

loss. It can be c l e a r l y seen from f i g u r e s 3 t h a t t he working f l u i d ( to luene )

s t i l l has cons iderable amount of energy a f t e r expansion i n t h e power c y l i n d e r ,

and t h a t a l a rge condenser is requi red t o br ing to luene back t o the l i q u i d

state. The ca l cu la t ed cyc le e f f i c i e n c y is in t h e 20% range f o r the bottoming

c y c l e and this agrees wi th previous estimates.

5. LTX(=BANGER ~SIDEPATXOHS

Design, cons t ruc t ion and i n s t a l l a t i o n of t he evapora tor poses t h e

b i g g e s t cha l lenge i n a practical a p p l i c a t i o n of t h e bottoming cyc le i n a

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10

t ruck . Even though t h e DOE demonstration t ruck provfded a s o l u t i o n , a

product ion ve r s ion should be much more compact and c o s t e f f e c t i v e . The

i n t e g r a t e d concept s tud ied here provides an oppor tuni ty t o make t h e bottoming

cyc le p a r t of t h e engine and, therefore , should be under the hood of t h e

t ruck. .The c r i t i ca l component t o make such an i n s t a l l a t i o n poss ib l e is t h e

evaporator . This s e c t i o n descr ibes a f i r s t c u t a t t h e des ign of t h e

evaporator .

For t h e purpose of ana lys i s , t h e evapora tor is d iv ided i n t o t h r e e

d i s t i n c t s e c t i o n s , namely t h e economizer, t he evapora to r and t h e

superhea ter . Since t h e source of heat f o r t h i s hea t exchange? is exhaust gas,

tube and f i n type hea t exchanger should be used with t h e exhaust gas passing

ove r t he f in s . The tube s i d e hea t t r a n s f e r Coef f i c i en t can be c a l c u l a t e d from

equat ion (1) .

,f . * c ' t - - .I NU = ~e $8 Pr A Eq. (1)

where :

t h e Reynold's number, R e = VD/i, , t he P r a n t l ' s number, Pr = cp/k and

t h e Nusse l t number, Nu = h i D/k

The hea t t r a n s f e r c o e f f i c i e n t on t h e exhaust gaa s i d e is bf t h e orderof

. 10 Btu/(hr-f&- deg. P). Since the tube s i d e hea t t r a n s f e r c o e f f i c i e n t s are

g e n e r a l l y two o r d e r s of magnitude greater than t h e f i n s i d e hea t t r a n s f e r

c o e f f i c i e n t s , t he o v e r a l l hea t t r a n s f e r c o e f f i c i e n t is e s s e n t i a l l y l i m i t e d by

t h e gas s ide . Reference 8 recommends a va lue of 5-6 Btu/(hr-f(l2- Deg. F) f o r

the o v e r a l l hea t t r a n s f e r c o e f f i c i e n t f o r a hea t exchanger wi th 1/8" t h i c k

-.

139

steel tube. A value of 5.5 was used i n t h i s study. Table 8 g ives the

ca l cu la t ed heat transfer su r face areas and box volumes f o r the evaporator f o r

t he var ious cases. Since the box volume 1s of g r e a t importance t o the engine

des igner , these r e s u l t s are also shown as graphs i n f i g u r e s 7-9 compared t o

t h e recommendation of a hea t exchanger manufacturer (Ref 9 ) should be followed

and the f i n spacing should be l imi t ed t o 6 f ins / inch .

An examination of f i g u r e s 7-9 i n d i c a t e s t h a t the evapora tor s i z e would

be smaller i n t h i s i n t e g r a t e d bottoming cyc le compared to the u n i t i n t he DOE

demonstration truck. However, t he a c t u a l s i z e is s t i l l too l a r g e f o r "under

the hood" i n s t a l l a t i o n unless some c l e v e r packaging is designed.

The condenser p a r t of t he bottoming cyc le 1s t h e same as t h e t ruck

r ad ia to r . S l i g h t l y superheated steam from the expander o u t l e t mixes wi th t h e

engine coolant a t o r j u s t before the top tank of t he r ad ia to r . Since t h e

pressures of the two streams are designed to be the same, it is expected t h a t

t h e mlxing process a lone w f l l condense t h e superheated steam i n t o atleast

p a r t i l l y s a tu ra t ed steam. This assumption f s reasonable due t o t h e f a c t t h a t

t h e mass flow rate of the working f l u i d is much smaller than t h e mass flow

r a t e of t he engine coolant . In orde r t o handle the e x t r a hea t load , i t is

es t imated tha t t h e r a d i a t o r will have to be enlarged by 25-302;. More d e t a i l e d

d e s c r i p t i o n of the methodology and economic a n a l y s i s of the bottoming c y c l e s

wi th va r ious working f l u i d s can be found i n re f . 10.

6. OTBgP DESIGN CONSIDEUTIOEJS

The major d i f f e r e n c e between t h i s s y s t e m and the DOE demonstrat ion

t r u c k system is t he power expander. It is proposed t h a t one of the s i x

140

12

a y l d d e r s of t h e engine be used as t h e power recovery device. Such a -< -

r e c i p r o c a t i n g expander has been s tudied before ( r e f . 6,111. Besides, steam

r e c i p r o c a t i n g engine concept is q u i t e o ld and w e l l known. However, when t h e

- --

r e c i p r o c a t o r is i n t e g r a t e d i n t o the d i e s e l engine, s e v e r a l des ign i s s u e s must

be addressed. The speed and power output of t h e expander must always be t h e

same as the rest of the cy l inders . I n a t r a n s p o r t a t i o n a p p l i c a t i o n t h i s might

pose a problem a t l i g h t loads , i d l e and low speeds when the bottoming cyc le is

not very e f f i c i e n t . The expander cyli 'her w i l l produce power every revolut ion-- . -

whereas the rest of t h e c y l i n d e r s of a Cummins engine ope ra t e on a f o u r s t r o k e

cycle. This means the cam and the valve systems f o r t h e expander must be

d i f f e r e n t than for t h e o t h e r f i v e cy l inders . The i n t a k e va lve should be open

very r a p i d l y t o a l low a l l the working f l u i d i n t o the c y l i n d e r qu ick ly and

s t i l l have adequate t i m e f o r expansion work. The exhaust event can be more

gradual . The displacement of the expander is determined by t h e "swallowing

capac i ty" needed t o accommodate t h e requi red mass of the working f l u i d a t t h e

lowest p re s su re in t he cycle . T h i s c a l c u l a t i o n is i l l u s t r a t e d i n Table 9 .

The same cam t h a t ope ra t e s t he valves and t h e i n j e c t o r s i n t h e f i v e d i e s e l

c y l i n d e r s can be used f o r t h e expander, i f t he requi red va lve even t s could be

accomplished by s imple modif icat ions such as double lobed, s teep ramp cam

p r o f i l e . Some innovat ive va lves such as s l i d i n g va lves should be cons idered

f o r t h e expander cy l inder . While these changes appear t o be d i f f i c u l t , t h i s

approach has t h e advantage of e l imina t ing the tu rb ine and the gear t r a i n . A

d e t a i l e d a n a l y s i s of t h e r ec ip roca t ing expander should be done in t h e next

phase of t he concept evaluat ion.

14 1

13

7. CONCLUSIONS

I . The integrated rankine bottoming cycle is a feasible

concept for truck applications. This concept presents

. ,. significant cost reduction and system simplification

potentials for the ORBC system which has been

previously demonstrated.

2. Evaporator size could be significantly reduced compared

to the ORBC system. But the size is still quite large

and requires innovative ideas for under the hood

installation.

3. Steam at 1000 p s i pressure is the first choice for the

working fluid and toluene at 500 psi is the second

choice. Evaluation of the suitability of 50- 50

ethylene glycol / water mixture as working fluid should

be more thoroughly investigated.

1 4 2

1 2.-

' '-1

- I - BgFeaencEs

. ' ' A',, L \

1 . Hoehne, J. L. and J. R. Werner, The Cumins advanced Turbocompound Diesel

Engine Evaluat ion, DOE/NASA/4936-2, Apr i l , 1983.

2. - Bpands, M.C., J.R. Werner, J.L. Hoehne and S. Kramer, Vehicle Testing of

C M n s Turbocompound Diesel Engine, SAE Paper No. 810073.

3. Sekar, R., and R. Kamo, Positive Displacement Compounding of a Heavy

Duty Diesel Engine". Fina l Report for NASA/DOE c o n t r a c t NO. DE-AC02

-78CS54936. NASA CR No. 168286, November 1983.

4. "Diesel Organic iZanki*rs Cycle Compound Engine For Long Haul Trucks,

Cont rac to r ' s F i n a l Report by Thenuo E lec t ron Corporation. Report No.

T34257-236-82, DOE/CS/52832-4, U.S. Department of Energy Contract No.

DE-AC02-76CS52832

5. DiBella, F.A., L.R.DiNanno and M.D. Koplow, Luboratory and on-Highmy

testing of Dieset Oaganic Rankine Compound Long-Haul Vehicle Engine, SAE

pape r No. 830122

6. Poul in , E., Steam Bottoming Cgcte for Adiabatic Diesel Engine" Foster-

Miller Associates, NASA Contract No. DEN3-300, NASA CR No. 168255, 1984.

7. Bai ley , M.M. Comparative Evaluation of Three Alternative Power Cycles

f o r Waste Heat ,?ecovery from the Exhaust of Adiabatic Diesel Engines,

NASA TM-86953, DOE/NASA/50194-43, J u l y 1985.

143

8. Boyen, J.L., John Wlley d Sons, h e . , Thermal Energy Recovery", 2nd New

York 1980

9 . 'Gayhart, E. and E. Ph i l l ips , A firchasers Guide to Diesel Engine Exhaust

Heat Recovery, Vapor Corporation, Chicago, I l l i n o i s , November 1985.

10. Marciniak, T.J. et a l . , Comparison of Rankine Cycte Power Syi3tem8:

Effects of Seven Working Fluids", Argonne National Laboratory, ANLICNSV-

TM-87, June 1981.

11. Engineering Research on Positive L%3placem%nt Ga8 Expanders. Technical

Report for DOE Contract No. DE-AC02-83CE40652 by Northern Research and

Engineering Corporation. Report No. 1514-2, February 1984

1 4 4

HIGtIBE 1 Schartfc Magram of tbc Organic Rankine Bottwng Cycle Demonstrated In 8 DOE Test Vebicle

(Figure adopted from Reference 4)

14 5

Steam a t 1000 PSI

Exhaust-> - - ( 5 ) +

*

L

I -+ Steam from t h e > I r ec ip roca t ing expander

I ( 6 ) \

Radiator Top Tank

(7) 1 Radiator

By-Pas s I 1 , -

(1) (2) Englee bad (3) Mala engine coalant p n p (4) Bottomlog cycle booster pamp (5) Evaporator (6) (7)

S i x cylinder tmck diesel engine

One cylfndcr of the engfne used as bottoming cycle expander Truck radiator, also condenser for the IRBC

FXGUBE 2 Integrated Rankine bottoming Cycle (IPIIC) Concept

146

ORIGINAC PAGE IS OF POOR QUALITY

A 0 0.1

147 FIGURE 3 T-S Magtam of the IRBC: S t e a at 1000 PSI as Vorklng P l d d

u(Ta0cY. .. B N A U m a

FIGURE 4 T-S M a g r a of I R X : Steam at 500 PSI as Worklug l l u i d 148 .

FIGURE 5 T-S D i a g r a m of IRBC: Toluene at 500 PSI with Regeneration

. .

. I

. . . . . . . . . . . 1 ' - - '

. .

. . + : : , . I "

. ' . 11100: - . . d . . . .

I " I " '

I . . . ' . .-,-

. . ; . . . .

. , . . .

. . . . . , . . . .

' ' 1 :,:-cZ . , . . . .

. . . . . . . . . . . . , . . . . . . . . . . . . . . .

. . . / . . . . . . . . . . . I . . . . . . . - - . . . . . I

. . . . . . . . . . . . ; . . , ___

- - . . . .

. . . . , ' ' _ ! .

. . . . 1 : . . ! . . . i . . . . : . : : : I Evapordtor _I - '

I I , . . . . .3. . . , . . ... +.. . . 0 , ... I . - - . . ' - 4 . 4 . - 8 - . . - . . . . . . : . I . . . . . . _ . I . .

FIGURE 6 I - S M y r u of IRBC: Toluene at500 PSI, No Regenerator

EVAPORATOR SIZE POR THE INTEGRATED W I N S - -_- BOITO)(ING CYCLE FOR TRUCK DIESEL ENGINE

1 U . L

8.0

6.0

4.0

2.0.

1

Sream @ 1000 PSI

0 I 1 1 I

... '. . '.

x. \

-f i n s / i nc h

f i n s / i n c h

fins / i n c h

Working Fluid Inlet Temperature to the Evaporator, O F

X - Department of energy Demonstration Truck Evaporator Desfgned 6 Bullt by

rheraoclect con

Evaporator S ize for t h e Integrated Rankine Bottodng Cycle for Truck Diesel Engine - Steam e 1000 PSI

151

I I I

P I C m 8 Evaporator Slze for LRBC: Steam at 500 PSI

Tube

O u t s i d e Diameter = 1 / 2 " or 5/8"

Wall t h i c k n e s s '= 0.04" t o 0.05"

Material = c a r b o n steel

p i t c h p = 3 / 4 "

F i n s

Th ickness = 0.015" t o 0.020"

Material = low c a r b o n s tee l

No. of f i n s / i n c h = 6

M u l t i p l e t u b e and f i n compact

Heat exchange d e s i g n i s recommended.

Brazed t u b e t o f i n j o i n t shou ld be used

O v e r a l l d imens ions t o s u i t packag ing on t h e e n g i n e

T o t a l F i n . s u r f a c e area should match t h e f i g u r e

chosen from t h e g raphs .

FIGURE 10 Evapora to r Tube and F i n D e t a i l s

1 5 3

. 4 Y

4 b

Y Y U

f 2 4 L - Y

0 4

E

0

I.

Y U

0

CI

5

4 Y . 4 LI 4 Y

C Y P CI

0 V

3,

Y P P r. -.

1

1

0

n H N

- A

s1

! 2

n x U

Y

P

Y

0

Y

I

Y 0

il

0

Y

0

Y s

N

h

N

N -

9 N

N n

51 0,

N

9 ‘D

I. Y Y .

\

0 V

P m

e Y

8 O

I - c Y P

c

U I.

U

c P

P

4

4 - I

2 m N

9 h -

Q,

0 3

n *

ij 0 m

- 2

t n

ON vi 0 U

. e m

8 0

0 u

Y

Y > a

0 hI.

* Y u m > c m 0

:

CI

0 -

d -

n n N

I

5 U

N - x

Y

e a Y

P

I d

ZGINAL PAGE I$ POOR QUGI;ITY

154

. Table 2 Ihermal scab l l l ty W e 8 of Thumb

A l i p h a t i c Compounds

a l c o h o l s

amines

ke tones

e t h e r s

a c i d s

esters

h yd roca r bo ns

c h l o r i d e s & f l u o r i d e s

s i l a n e s & si l icates

per f luorocarbons

p e r f l u o r o e t h e r s

perf luoroamines

Aromatic Compounds

hydrocarbons

e t h e r s , amines, s u l f i d e s

c h l o r i d e s

f l u o r i d e e

Generalized S t r u c t u r e

RCHZOH

RCHZOCHZR 0

0 R ~ H

R & O R ~

RCH2X

ArH, ArCH3, At2CH2

A r O r , A r 3 N , Ar2S

Arc1

ArF

Approximate Decomposi t i o n Temperature, OF

250 - 650

250 - 650

250 - 450

500 - 600

200 - 600

350 = 615

630 = 680

200 - 500

580 - 680

800 - 900

750 - 850

750 - 850

800 - 1000

850 = 900

700 = 800

800 - 850

1 5 5

+ Table 3 Deflnitloas of Toxicity

Toxicity

Approximate lethal oral dose for a 70 kg man tD50

none >15 g/kg > 1 quart

slight 5-15 g/kg 1 quart

moderate 0.5-5 g/kg 1 pint

high 50-500 %/kg 1 ounce

serious 1-50 mg/kg 1 teaspoonful

dangerous <1 mg/kg a taste

156

Table 4 Definitione of Fire and Explosion Hazard

Fire and Flash point exploelon hazard

<lOO°F dangerous

100-200.F moderate

>200°F low

c

157

Table 5 Additional Fire Eazard Data

Autoignition Fluid Flash hint O F Fire Point O F Temperature OF

RC-1 none none none

Pluoriaol 85 105 160

Z-aethylpyridine/water 130 145 1060

toluene 40 900

12 928-1044

54-55 727-878 100% met hano 1

100% ethylene glycol 232 748-752

kerosene 100-1 60

gasoline -50 536-8 53

* benzene

diesel X2 100 494

-

158

TABLE 6

Steam Rankine Cycle Expander Cycle Calculations (An Example)

State Point

1

2

3

4

5

6'

6

7

T

2 50

- 545

545

950

2 50

373

2 50

P

29.82

1000

1000

1000

1000

29.82

29 . a2

29 .82

Notes a) T = Temperature, OF

p = Absolute pressure, psia

v = spec i f i c volume, c . f t l l b

h = enthalpy, Btu/lb

V

.017006

- .02159

.44596

.7953

13.128

16.440

i 3 . a i 9

h

218.59

222.95

542.60

1192.90

1477.10

1116.67

1224.80

1164.00

c) pump ef f ic iency , rl pump - 0.7

expander ef f ic iency neXp = 0.7

159

TABLE 6 (cont . )

Sample Calcu la t ion

Steam Rankine Cycle - Steam a t 1000 p s i

(h -h ) - (h2-hl)

h5'h2 Cycle e f i c i e n c y = 5 6

(1477-1255) - (223 - 219) - 1477

= 0.198

Energy input = h c AT of exhaust gas P

= SO x 0.25 (110-500)

= 7500 Btu/min

Power output = 7500 x 0.198 42.5

7500 Theoretical steam flow r a t e = - h -h 2 2

= 7500 1477-223

= 5.98 lb/min

26 lb/min

160

T U L E 7 Reeults of Cycle Ana lys i s for t h e Different Cases

Work f ng Flu id

Expander Operating C y c l e F l o w Rate P o w e r Pressure E f f i c i e n c y lb,/rnln H e P.

I . Superheated Steam 1000 19.8 5.98 34.9

2. Superheated Steam 500 19.5 6.10 34.4

3. Superheated Toluene w i t h Regeneration 500 18.0 32.40 32.1

4. Superheated Toluene wi thout Regeneration 500 13.6 24 30 24.1

161

rABLE 8 Heat Exchauger Site Calculation (Example)

1. Liquid i n l e t temp. O F 200 250 400

2 . Steam o u f t l e t temp. O F 950 950 950

3 . Exh. g a s i n l e t temp. O F 1100 1100 1100

4 . I n l e t temp d f f € . "F 470 500 600

m c n qnn 5. I n l e t temp d l t t . "i; 2 7B L >u LUU

6 . Out le t temp d iEf . O F 150 150 150

7 . *LMTD O F 2 04 196 174 .

8. Heat t r a n s f e r r a t e Btu/min 7876 741 1 6211

9. Overal l heat t r a n s f e r c o e f f i c i e n t . Btu lhr - f t 2

10. Heat t r a n s f e r area, f t 2 386 378 357

€ t 2 55 55 5 s - 1 1 . Surface area

Core Volume f L 3

12. Heat exchanger core volume, f t 3 7 - 0 2 6.87 6 .49

*Long-MEAN temperature d i f f e r e n c e

162

TABLI 9 Calculation of Expander Displacement

S p e c i f i c volume of steam at exapnder o u t l e t - 16.44 C. f t l b

Steam f low rate 6 lb /min .

= 6 x 16.44 c . f t 98.64 c . f t - min mln

If the engine is reated a t 1900 tpm, Steam flow rate - 98.64 c . f t 0.0519 toft

1900 rev r e v .

One exh. s t roke /rev . displacement * 0.0519 - c . f t s t r o k e

- 89.68 c . i n 90 c o i n V o l . 0.9, Displacement = 100 c . f n

163

APPENDIX 5

ADDITIONAL COST FOR INTEGRAL

BOTTOMING CYCLE ENGINE

vs . L10 TURBOCHARGED AFTERCOOLED ENGINE

165

ORIGTNAI; PAGE IS OF POOR QUACITY AOOITIONAL COST FOR INTEGRAL BOTTCWNG CYCLE ENGINE

W L 10 TURUTHUQO AFTERQIXEO ENGINE 0ATE:DECEMOER 17, 1986

166

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167

Nalmnal Aeronautics and

1 . Report No.

NASA CR-180833

Report Documentation Page 2. Government Accession No. 3. Recipient's Catalog No.

----- ------ 4. Title and Subtitle

Technical and Economic Study of S t i r l i n g and Rankine Cycle Bottoming Systems f o r Heavy Truck Diesel Engines

5. Report Date

September 1987 6. Performing Organization Code

------

9. Performing Organization Name and Address

Cumins Engine Company, Inc. Box 3005 Columbus, I N 47202-3005

7. Author@)

I. Kubo

11. Contract or Grant No.

DEN 3-361

8. Performing Organization Report No.

CTR 0723-87001

10. Work Unit No. -------

I 13. Type of Report and Period Covered

17. Key Words (Suggested by Author@))

2. Sponsoring Agency Name and Address

U.S . Department of Energy Off ice of Vehicle and Engine R €4 D Washington, DC 20585

18. Distribution Statement

Contractor Report

DOE/NASA/0361-1

19. Security Classif. (of this report) 20. Security Classif. (of this page) 21. No Of Pages Unclassif ied Unclassif ied

5. Supplementary Notes Fina l Report. Prepared under Interagency Agreement DE-AIO1-86CE50162 Pro jec t Manager, M. B a i l e y , Propulsion Systems Division, NASA L e w i s Research Center, Cleveland, OH 44135

22. Price'

----

6. Abstract For t h e evaluation of bottoming cycle concepts on heavy duty t r anspor t engine

1. 2 . Life-cycle c o s t eva lua t ion of t h ree bottoming systems: Organic Rankine,

3. Variables considered fo r t he second t a s k w e r e i n i t i a l c a p i t a l investments, fue:

savings, depreciat ion tax b e n e f i t s , salvage va lues , and service/maintenance c o s t s . The s tudy shows t h a t n o n e o f the th ree bottoming systems s tudied a r e even marginall! a t t r a c t i v e . Manufacturing costs have to be reduced by a t least 65%.

app l i ca t ions , following t a s k s were performed: Develop conceptual design and c o s t da t a f o r S t i r l i n g systems.

Steam Rankine, and S t i r l i n g cyc les . Suggest f u t u r e d i r e c t i o n s i n waste hea t u t i l i z a t i o n research.

AS a new approach, an in t eg ra t ed Rankine/Diesel system w a s proposed. I t u t i l i z e s one of d i e s e l cy l inde r s as an expander and c a p i t a l i z e s the hea t energy t o the engine coolant. The concept e l imina tes t h e need f o r the p o w e r transmission device and a sophis t ica ted con t ro l system, and reduces t h e s i z e of t h e exhaust evaporator . The system would o f f e r an a t t r a c t i v e package f o r end-users, giving roughly a 20% IRR a t a $1.25/gallon f u e l p r i ce . Further opt imizat ion of t he syster i s poss ib l e by eliminating/combining sane of concepts b u i l t i n t he cu r ren t design.

I I

'For sale by the National Technical Information Service, Springfield, Virginia 22161 NASA FORM 1620 OCT 80

~~

Waste hea t recovery S t i r l i n g cyc le Rankine cyc le

Unclassified-unlimited Star Category 85 DOE Category UC-96

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