EPA-460/3-74-013-b
July 1974
CURRENT STATUS
OF
ALTERNATIVE AUTOMOTIVE
POWER SYSTEMS
AND FUELS
VOLUME II -
ALTERNATIVE AUTOMOTIVE
ENGINES
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Waste Management
Office of Mobile Source Air Pollution Control
Alternative Automotive Power Systems Division
Ann Arbor, Michigan 48105
-------
EPA-460/3-74-013-B
CURRENT STATUS
OF
ALTERNATIVE AUTOMOTIVE
POWER SYSTEMS
AND FUELS
VOLUME II -
ALTERNATIVE AUTOMOTIVE
ENGINES
Prepared by
The Environmental Programs Group
The Aerospace Corporation
El Segundo, California 90245
Contract No. 68-01-0417
EPA Project Officer: Graham Hagey
Prepared for
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Air and Waste Management
Office of Mobile Source Air Pollution Control
Alternative Automotive Power Systems Division
Ann Arbor, Michigan 48105
July 1974
-------
This report is issued by the Environmental Protection Agency, to report
technical data of interest to a limited number of readers. Copies of this
report are available free of charge to Federal employees, current contractors
and grantees, and non-profit organizations - as supplies permit - from the
Air Pollution Technical Information Center, Environmental Protection Agency,
Research Triangle Park, North Carolina 27711, or may be obtained, for a
nominal cost, from the National Technical Information Service, 5285 Port
Royal Road, Springfield, Virginia 22151.
This report was furnished to the U.S. Environmental Protection Agency
by the Aerospace Corporation, El Segundo, California, in fulfillment of
Contract No. 68-01-0417 and has been reviewed and approved for publication
by the Environmental Protection Agency. Approval does not signify that
the contents necessarily reflect the views and policies of the agency.
The material presented in this report may be based on an extrapolation of
the "State-of-the-art" . Each assumption must be carefully analyzed by
the reader to assure that it is acceptable for his purpose. Results and
conclusions should be viewed correspondingly. Mention of trade names
or commercial products does not constitute endorsement or recommendation
for use.
Publication No. EPA-460/3-74-013~b
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FOREWORD
This report, prepared by The Aerospace Corporation for the
Environmental Protection Agency (EPA), Alternative Automotive Power
Systems Division, summarizes available nonproprietary information on the
technological status of automotive power systems which are alternatives to
the conventional internal combustion engine, and the technological status of
nonpetroleum-based fuels derived from domestic sources which may have
application to future automotive vehicles.
The status of the technology reported herein is that existing at
the end of 1973 with more recent data in selected areas. The material pre-
sented is based principally upon the results of research and technology
activities sponsored under the Alternative Automotive Power Systems (AAPS)
Program which was originated in 1970 and which is administered by the Alter-
native Automotive Power Systems Division of EPA. Supplementary data are
included from programs sponsored by other government agencies and by pri-
vate industry. Additional information on technology and development programs
is known to the government but cannot be documented herein because the data
are proprietary.
One purpose that the AAPS Program serves is to provide a basis
of knowledge and perspective on what can and cannot be accomplished with
the use of alternative propulsion and fuels technology and to disseminate this
information to Congress, Federal policy makers, industry, and the public.
Thus, the publication of information such as that contained herein is in keeping
with this element of the mission of the AAPS Program. This is the first of a
series of reports on alternatives that are intended to be published annually.
The results of this study are presented in four volumes and three
main topical areas:
Volume I. Executive Summary
Volume II. Alternative Automotive Engines
111
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Volume III. Alternative Nonpetroleum-based Automotive Fuels
Volume IV. Electric and Hybrid Power Systems
Volume I, the Executive Summary, presents a concise review of important
findings and conclusions for all three topical areas. Thus, an overview of
the study results may be obtained by reading Volume I only. Volumes II,
III, and IV contain detailed, comprehensive discussions of each topical area
and are therefore of interest primarily to the technical specialist. Each of
these three volumes also contains Highlights and Summary sections pertaining
to the topical area covered in the volume.
This volume, Volume II, presents available information pertaining to
the current technological status of advanced alternative automotive heat engines
which may have application to future automotive vehicles.
A brief review of important findings and conclusions is presented in
the Highlights and Summary sections. A short discussion of the background of
the Alternative Automotive Power Systems (AAPS) Program sponsored by EPA
is given in Section 1, Introduction. Section 2 presents a brief historical
review of the evolution of the spark ignition engine powered automobile and
discusses automobile systems and operational requirements with respect to
the interrelationships influencing performance capability, fuel economy, and
exhaust emissions as characterized by current Federal exhaust emissions
standards. Sections 3 through 8 address in detail the available data concern-
ing the six alternative engine classes examined: gas turbines, Rankine cycle
engines, Stirling engines, diesel engines, Wankel and other rotary piston
engines, and stratified charge engines. The principal topics covered for each
engine class include power plant description, performance characteristics,
and current and projected technology status. Appendices A and B contain a
listing of vehicle specifications for full-size and compact passenger cars.
These specifications were developed in the AAPS Program to provide a refer-
ence basis for comparison of vehicle performance when powered by various
alternative engines.
IV
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ACKNOWLEDGMENTS
Appreciation is acknowledged for the guidance and assistance
provided by Mr. Graham Hagey of the Environmental Protection Agency,
Alternative Automotive Power Systems (AAPS) Division, who served as
EPA Project Officer for this study. Appreciation is also extended to staff
members in the AAPS Division, to other EPA divisions, to various govern-
ment agencies and to those in industry and the academic community who
supplied reference material and reviewed the contents of this report.
The following technical personnel of The Aerospace Corpora-
tion made valuable contributions to the effort performed under this contract:
Owen Dykema
Lester Forrest
Farhad Ghahremani
Richard Kopa
Warner Lee
Wolfgang Roessler
William Smalley
Merrill G. Hinton, Director
Office of Mobile Source Pollution
D. E. Lapedes,/Study Manager
To"ru lura, Associate Group Director
Environmental Programs Group
Directorate
JoVeph ivLeltzer, Group director
//^Environmental Programs Group
Directorate
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CONTENTS
FOREWORD iii
ACKNOWLEDGMENTS v
HIGHLIGHTS H- 1
SUMMARY S-l
1. INTRODUCTION TO HEAT ENGINE SECTION 1-1
2. OPERATIONAL REQUIREMENTS AND CHARACTERIS-
TICS OF AUTOMOBILES POWERED BY CONVENTIONAL
SPARK IGNITION ENGINE 2-1
2. 1 Evolution of the Spark Ignition Engine 2-1
2.2 Automobile Systems and Operational Requirements:
Interrelationships Influencing Performance and
Fuel Economy 2-6
2.3 Exhaust Emissions and Emission Control Systems .... 2-20
3. GAS TURBINE ENGINES 3-1
3.1 Introduction 3-1
3.2 Power Plant Description 3-4
3.3 Performance Characteristics 3-33
3.4 Current Status of Technology 3-41
3.5 Projected Status 3-62
4. RANKINE CYCLE ENGINES 4-1
4. 1 Introduction 4-1
4.2 Power Plant Description 4-3
4.3 Performance Characteristics 4-17
4.4 Current Status of Technology 4-19
4. 5 Projected Status 4-50
Vll
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CONTENTS (Continued)
5. STIRLING ENGINES 5-1
5. 1 Introduction 5-1
5.2 Power Plant Description 5-3
5.3 Performance Characteristics 5-28
5.4 Current Status of Technology 5-33
5.5 Projected Status 5-34
6. DIESEL ENGINES 6-1
6. 1 Introduction 6-1
6.2 Power Plant Description 6-2
6.3 Performance Characteristics 6-24
6,4 Current Status of Technology 6-55
6. 5 Project Status 6-56
7. WANKEL AND OTHER ROTARY PISTON ENGINES 7-1
7.1 Introduction 7-1
7.2 Power Plant Description 7-3
7.3 Performance Characteristics 7-15
7.4 Current Status of Technology 7-18
7.5 Projected Status 7-19
8, STRATIFIED CHARGE ENGINES 8-1
8. 1 Introduction 8-1
8.2 Open-Chamber Stratified Charge Engines 8-3
8.3 Divided-Chamber Stratified Charge Engines 8-46
8.4 Concluding Remarks 8-75
Vlll
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CONTENTS (Continued)
APPENDIX A: ENVIRONMENTAL PROTECTION AGENCY,
PROTOTYPE VEHICLE SPECIFICATION,
FULL-SIZE PASSENGER CAR
A-l
APPENDIX B: ATTACHMENT, ENVIRONMENTAL PROTECTION
AGENCY, PRELIMINARY PROTOTYPE VEHICLE
SPECIFICATION, COMPACT PASSENGER
CAR
ABBREVIATIONS
GLOSSARY ....
REFERENCES .
B-l
Ab-1
Gl-1
R-l
IX
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FIGURES
2-1. Generalized Engine Torque Characteristic 2-8
2-2. Thermal Efficiency over the Federal Driving Cycle 2-10
2-3. Performance Map of a Carbureted Spark Ignition Engine 2-11
2-4. Generalized Engine and Vehicle Torque
Characteristics 2-12
2-5. Fuel Economy over the Federal Driving Cycle 2-15
2-6. Motive Power Requirements, 4,300-pound Vehicle 2-17
2-7. Warmup Economy 2-19
2-8. Effect of Air Fuel Ratio on Emission Levels, Gasoline
Spark Ignition Engine 2-21
2-9. General Motors Projected 1975 Under-Floor
Emission Control System 2-25
3-1. Simple Gas Turbine Engine, Schematic Diagram 3-2
3-2. Free-Turbine Engine, Schematic Diagram 3-2
3-3. Regenerative Free-Turbine Engine, Schematic
Diagram -^"^
3-4. Temperature-Entropy Diagram for Simple Cycle
Gas Turbine
3-5. Equipment for Gas Turbine Regenerative Cycle,
Schematic Diagram ............................. 3-7
3-6. Estimated Turbine Performance, Single-Shaft
Recuperated Engine ............................. 3-11
3-7. Flow Path for Reverse-Flow, Can-Type
Combustors .................................. 3-13
3-8. Estimated Flame Temperatures and Equilibrium
NO Concentrations as Functions of Air Fuel
Ratio for Various Combustor Inlet Temperatures ......... 3-14
XI
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FIGURES (Continued)
3-9. Disc-Type Rotary Regenerator 3-15
3-10. Estimated Effect of Pressure Ratio and Turbine Inlet
Temperature on Fuel Economy for Nonregenerative
Cycle 3-23
3-11. Estimated Effect of Pressure Ratio and Turbine Inlet
Temperature on Fuel Economy for Regenerative
Cycle : 3-23
3-12. Typical Fixed and Free Turbine Torque and Power
Characteristics 3-25
3-13. Estimated Road-Load Fuel Economy of Single-Shaft
Regenerated Engines, 85°F Sea-Level Day 3-27
3-14. Engine Braking Force Characteristics 3-29
3-15. Control System Functional Schematic, Single-Shaft
Gas Turbine Engine 3-31
3-16. Standing-Start Acceleration Performance 3-40
3-17. Starting Time versus Ambient Temperature 3-42
3-18. Schematic of Chrysler Sixth-Generation, Gas
Turbine Engine 3-46
3-19. Chrysler 150-hp Gas Turbine Engine 3-47
3-20. Chrysler Baseline Turbine-Powered Vehicle 3-49
3-21. Chrysler 150-hp Gas Turbine Engine,
Vehicle Installation 3-50
3-22. Chrysler 100-hp Upgraded Gas Turbine Engine with
Single Vertical Regenerator Compact Vehicle
Installation 3-51
3-23. Chrysler 100-hp Upgraded Gas Turbine Engine with
Single-Shaft Vertical Regenerator Compact Vehicle
Installation 3-52
3-24. Solar Jet-Induced Circulation Combustor, JIC-B 3-55
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FIGURES (Continued)
3-25. Schematic of Aerojet Platelet Premixer Concept 3-56
3-26. AiResearch Recuperator Bypass Combustor 3-58
3-27. General Electric Low-NOx Porous-Plate
Gas Turbine Combustor 3-59
4-1. Typical Rankine Cycle System, Schematic 4-2
4-2. Typical Ideal Rankine Thermodynamic Cycle -with
Water as a Working Fluid, Schematic 4-6
4-3. 150-Horsepower Steam Engine 4-21
4-4. Steam Reciprocating Engine 4-22
4-5. Organic Reciprocating Engine 4-23
4-6. Organic Turbine Engine 4-24
4-7. Steam Turbine Engine 4-25
4-8. Prototype Installation, Condenser Removed,
Plymouth "C" Body 4-26
4-9. Prototype Installation, Plymouth "C" Body,
Top Section 4-27
4-10. Prototype Installation, Plymouth "C" Body,
Side Section 4~28
4-11. Current and Projected Fuel Economy for Rankine
Cycle Engines 4-32
5-1. Basic Stirling Displacer Engine, Schematic 5~2
5-2. Stirling Displacer Engine with Rhombic Drive, Schematic. ... 5-2
5-3. Thermodynamic Description of the Stirling Cycle 5-4
5-4. Stirling Engine Efficiency versus Specific Power Output 5-7
5_5. Displacer Type Engine, Schematic 5-8
Xlll
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FIGURES (Continued)
5-6. 1973 Ford Torino with Philips Stirling Double-Acting Engine. . 5-10
5-7. Philips Stirling Double-Acting Engine with Swash-Plate Drive . 5-11
5-8. Parts Layout for Philips Stirling Double-
Acting Engine, Torino Application 5-12
5-9. Front 3/4 View of Philips Stirling Double-
Acting Engine Mock-Up for a Ford Torino 5-13
5-10. Double-Acting Type Engine, Schematic 5-15
5-11. Double-Acting Piston 5'16
5-12. Double-Acting V-4 Engine, Preprototype 5-17
5-13. Double-Acting Cylindrical Configuration 5-18
5-14. Dynamomenter Torque of a 40-hp Stirling Engine as a Function
of the Speed r\ at Different Values of P max 5-19
5-15. Schematic of Pressure-Throttling Method 5-25
5-16. Power Control Diagram 5-26
5-17. Fuel Control Diagram 5'27
5-18. Performance Map of Ford Torino Stirling Engines 5-29
6-1. Ideal Diesel Cycle 6'4
6-2. Diesel and Otto Cycle Thermal Efficiency vs
Compression Ratio
6-3. Low-Swirl Diesel Engine °
6-4. Medium-Swirl Diesel Engine 6-8
6-5. High-Swirl, "M" System Diesel Engine 6-8
6-6. Ricardo Swirl Chamber Diesel 6-9
6-7. Daimler-Benz Precombustion Chamber Design 6-9
xiv
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FIGURES (Continued)
6-8. Air-Ceil Diesel Engine Schematic .................... 6-10
6-9. Specific Power Output of Diesel and Gasoline Engines ...... 6-12
6-10. Comprex Supercharger Arrangement .................. 6-14
6-11. Specific Weight of Diesel and Spark Ignition Engines ....... 6-16
6-12. Specific Volume of Diesel and Spark Ignition Engines ....... 6-18
6-13. Diesel and Gasoline Engine Performance Parameters,
Daimler-Benz Automobile Engines .................... 6-25
6-14. Precombustion Chamber Diesel and Uncontrolled
Spark Ignition Engine Performance Maps ............. . . 6-27
6-15. Four-Cycle Diesel Engines ........................ 6-31
6-16. Exhaust Smoke vs Output Power for Four Open -Chamber,
Naturally Aspirated, Heavy-Duty Diesel Engines .......... 6-37
6-17. Effect of Exhaust Gas Recirculation on the Emissions
from a Light-Duty Diesel Engine .................... 6-40
6-18. Effect of Water Induction Rate on the NO Emissions
of a Turbocharged Diesel Engine .................... 6-41
6-19. Fuel Economy versus Vehicle Speed for Diesel and
Gasoline Automobiles ............................ 6-43
6-20. Fuel Economy versus Engine Rated Power, 1973 Model
Year Certification Vehicles ........................ 6-46
6-21. Interior Vehicle Noise, Mercedes Benz 220D Diesel
and 220 Gasoline Automobiles ...................... 6~48
6-22. Compression and Ignition Temperatures versus
Compression Ratio .............................. 6-58
7-1. Wankel Rotary Engine ............................ 7-2
7-2. Production Rotary Piston Engines for Automotive
Application, Cutaway Drawing ...................... 7-6
xv
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FIGURES (Continued)
7-3. Motion of Rotor in Main Housing of Rotary Piston
Engine 7-7
7-4. Variation in Size and Compression Ratio for Engines
with Fixed Swept Volume and Various K-Factors 7-9
7-5. Comparison of Rotating Parts 7-11
7-6. Silhouette Comparison of a Twin-Rotor Wankel Engine
with a 283-cubic inch Chevrolet V-8 Piston Engine 7-13
7-7. Comparison of Main Parts for RX-2 Rotary
Engine and 6-Cylinder Reciprocating Engine (-14
7-8. Typical Apex Seal 7-18
8-1. Baudry Stratified Charge Engine g-5
o f
8-2. Borg Warner Combustion System °
Q T
8-3. Hesselmann Stratified Charge Engine
8-4. Mitsubishi Combustion Process 8-8
8-5. Witzky Stratified Charge Engine 8-9
8-6. Full Throttle Performance of Witzky Stratified Charge
Engine, Conventional Gasoline Engine, and Diesel Engine. • . • 8-10
8-7. Texaco Controlled Combustion System 8-11
8-8. Texaco Cup Combustion Chamber^ Stratified
Charge Engine 8"12
8-9. Ford L-141 PROCO Engine 8-18
8-10. Ford Fast Burn Phase II Stratified Charge Engine 8-22
8-11. Full-Load Brake Performance, Texaco Naturally
Aspirated L-141 TCCS Engine, Gasoline 8-26
8-12. Multifuel Brake Performance, Texaco Naturally
Aspirated L-141 TCCS Engine 8-2?
xvi
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FIGURES (Continued)
8-13. Effect of Stratified Charge Combustion on Emissions 8-29
8-14. Level-Road Fuel Economy, M-151 Light-Duty Vehicle,
Texaco Naturally Aspirated, L-141 TCCS Engine 8-41
8-15. Ford Torch Ignition Engine 8-50
8-16. Jozlin Prechamber Concept 8-52
8-17. Bishop Prechamber Concept 8-52
8-18. Honda CVCC Divided Chamber Stratified Charge Engine 8-54
8-19. Broderson-Conta Engine 8-56
8-20. Heintz Ram Straticharge Engine 8-57
8-21. Nilov Engine 8-57
8-22. Volkswagen Prechamber Engine 8-59
8-23. Newhall Prechamber Engine 8-60
8-24. Phillips Single-Cylinder Prechamber Engine Emissions
and Indicated Specific Fuel Consumption vs Indicated
Mean Effective Pressure 8-67
8-25. Emission Characteristics, Volkswagen Single-Cylinder
Prechamber Engine 8-69
8-26. Preliminary Emission Data, Newhall Prechamber
Engine 8-70
xvn
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TABLES
2-1. Typical Load-Power Requirements 2-13
2-2. Three-Percent Grade Power Requirements for 4,300-pound
Vehicle 2-14
2-3. Auto Trip Statistics 2-18
2-4. Summary of Passenger Car Emission Standards 2-23
3-1. Engine/Vehicle Basic Characteristics 3-8
3-2. Engine Material Content and Weight 3-19
3-3. Estimated High-Temperature Metal Element Require-
ments for 10 Million Engines 3-20
3-4. Exhaust Emission Data over the Federal Driving Cycle 3-34
3-5. United Aircraft Estimated Fuel Economy. Baseline
Vehicles 3-36
3-6. AiResearch Estimated Fuel Economy of Free-Turbine
Regenerated Cycle (4, 000-lb Test Weight Vehicle, 4-hp 3-36
Accessory Load, Three-Speed Automatic Transmission)
3-7. Specifications of Chrysler Corporation's Fourth- 3-43
Generation Gas Turbine Engine
4-1. Summary of Rankine Cycle Engine/Vehicle Development 4-4
Program
4-2. AAPS Rankine Cycle Engine Program, Preprototype 4-29
Power Plant Descriptions
4-3. Preliminary Emission Results on Federal Driving Cycle .... 4-31
4-4 California Steam Bus-Project, Power Plant Description .... 4-36
4-5. California Steam Bus Project, Selected Power Plant
Specifications „ 4-37
4-6. Comparison of Emissions from Steam- and Diesel-Powered
Buses 4-38
5_1. CVS Test Simulation Emissions in gm/mile 5-30
xix
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TABLES (Continued)
5-2. Fuel Economy Comparisons ........................ 5-31
5-3. Performance Projections .......................... 5-33
6-1. Cetane Numbers and Autoignition Temperatures of
Different Fuels ................................ 6-20
6-2. Heavy-Duty Diesel Engine Emissions at Rated Power ....... 6-30
6-3. Average Light-Duty Diesel Vehicle Emissions ............ 6-34
6-4. Diesel Automobile Emissions ....................... 6-35
6-5. Smoke Emissions from Mercedes Benz 220D Automobile,
Simulated Federal Smoke Tests ..................... 6-38
6-6. Particulate Emissions from Mercedes Benz 220D Auto-
mobile, Dow Chemical Procedure, gm/mile ............ 6-38
6-7. Light-Duty Diesel and Gasoline Vehicle Fuel Economy,
miles /gallon .................................. 6-44
6-8. Exterior Vehicle Noise, Mercedes Benz 220D Diesel
and 220 Gasoline Automobiles ...................... 6-47
6-9. Interior Noise, Peugeot 504 Diesel and Gasoline
Automobiles .................................. 6-48
6-10. Average Odor and Gaseous Emissions from Mercedes
Benz 220D Automobile ........................... 6'51
7-1. Vehicles Equipped with Rotary Piston Engines ........... 7'4
7-2. Rotary Engine Developers ......................... '"•'
7-3. Best Emissions Results, Thermal Reactor and Wankel
Engine, 2750-lb Compact Car ...................... 7'16
7-4. Emissions at Low -Mile age Rotary Engine with
Oxidation Catalyst .............................. 7-17
8-1. L-141 Engine Specifications, Standard and TCCS .......... 8'15
8-2. PROCO Specifications, L-141 Engine ................. 8-19
xx
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TABLES (Continued)
8-3. Exhaust Emissions of a TUrbocharged M-151
TCCS Vehicle 8-31
8-4. Exhaust Emissions, Controlled Naturally Aspirated
M-151 TCCS Vehicle Equipped with Exhaust Gas
Recirculation System and Oxidation Catalysts 8-32
8-5. Effect of Emission Control System Modifications on M-151
TCCS Vehicle Emissions and Fuel Economy, 1975 FTP 8-33
8-6. Exhaust Emissions from a Chrysler Cricket TCCS
Vehicle, 2, 500-lb Inertia Weight 8-35
8-7. Emissions and Fuel Economy, M-151 PROCO Vehicles,
1975 CVS Test Procedure 8-36
8-8. 351-CID PROCO Durability--Vehicle: 1972 Montego,
110T722, 1975 Test Procedure 8-37
8-9. 351-CID PROCO Emissions and Fuel Economy--Vehicle:
1972 Montego, 110T722, 1975 Test Procedure 8-38
8-10. Vehicle Emissions and Fuel Economy, 351-CID Fast
Burn Phase I Engine in a 1972 Torino, 1975 Test
Procedure, Average of Two Tests 8-38
8-11. Vehicle Emissions and Fuel Economy, 400-CID Fast
Burn Phase II Engine in a 1973 Torino, 1975 Test
Procedure 8-39
8-12. Emissions and Fuel Economy, Honda CVCC
Vehicles 8-62
8-13. Low-Mileage Emissions and Fuel Economy of Vega
Vehicles Modified for CVCC 8-64
8-14. Low-Mileage 1975 FTP Emissions and Fuel Economy
of Impala Vehicles Modified for CVCC 8-64
8-15. Hot-Start Emissions and Fuel Economy, Ford Gran
Torino with Torch Ignition Engine 8-66
8-16. Ford Vehicle Hot-Start Emissions and Fuel Economy,
Three-Valve Prechamber Engines, 1975 FTP (Hot) 8-66
xxi
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TABLES (Continued)
8-17. Ford Large-Prechamber Emissions and Fuel Economy,
Single-Cylinder Engine 8-67
8-18. Current Prechamber Engine Development Status 8-73
xxii
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HIGHLIGHTS
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HIGHLIGHTS
Numerous power system concepts have been proposed for road
vehicles. In the category of heat engines for automobiles, several different
types of engines have demonstrated the ability to provide the desired power
levels while simultaneously offering the potential for major reductions in
exhaust emissions when contrasted with the conventional spark ignition
engine. In some cases, an engine has been characterized as having the
potential to meet the original 1975 Federal emission standards (hydrocarbon
= 0.41 grams per mile (gm/mi), carbon monoxide = 3.4 gm/mi, nitrogen
oxide = 3. 1 gm/mi) or the original 1976 standards (hydrocarbon = 0.41 gm/mi,
carbon monoxide = 3.4 gm/mi, nitrogen oxide = 0.4 gm/mi) without exten-
sive after treatment or add-on devices. Each engine, however, has certain
design and performance deficiencies to be overcome (viz. , excessive size,
high manufacturing cost, prolonged startup time, etc.). Both industry and
government have sponsored programs for overcoming these deficiencies
through a combined application of analytical and test efforts. Recent con-
cern with conservation of natural resources has prompted additional effort
in reducing engine fuel consumption and verifying flexibility of engine opera-
tion using a wide range of possible fuels. This effort is exemplified by
modifications within EPA's Alternative Automotive Power Systems (AAPS)
Program to stress reducing fuel consumption as well as achieving emission
levels consistent with Federal standards.
The following highlights briefly summarize the status of develop-
ment work on alternative engines investigated in recent years. The engine
types under review are: gas turbine, Rankine cycle, Stirling, diesel,
Wankel (and other rotary piston designs), and stratified charge. These
engine types are in various stages of development, ranging from redesign of
engines under limited production to extensive research and development
efforts on engines at least eight to nine years from full production (if placed
in a production program status).
H-l
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GAS TURBINE ENGINES
1. Until recently, the exhaust emissions of automotive-type gas turbine
engines have been typified as characteristically low in unburned hydro-
carbons and carbon monoxide, but with oxides of nitrogen levels too
high to permit attainment of the original 1976 Federal emission
standard of 0.4 gm/mi. For example, Chrysler's sixth-generation
engine installed in an intermediate-size vehicle has NOX emissions of
about 2.2 gm/mi over the Federal Emissions Test Driving Cycle.
2. However, within the AAPS Program, Solar, a Division of International
Harvester, has recently developed a new gas turbine combustion sys-
tem that, based on steady-state test data, is calculated to be capable
of meeting the original 1976 Federal emission standards and is close
to meeting the AAPS Program goals (one-half the level of the original
standards). Additional development is required to perfect the com-
bustor control system and to validate meeting program goals with
transient operation of the engine over the Federal Emissions Test
Driving Cycle.
3. Reflecting similar recent advancements, a General Motors-sponsored
program resulted in an experimental 225-horsepower gas turbine
engine installed in a car meeting the original 1976 Federal emission
standards. Complete details concerning engine design and operation
are not available.
4. Estimated high production costs continue to be a major factor inhibit-
ing the implementation of the automotive gas turbine engine. Other
problems involve the need for improvements in noise level, durability,
acceleration lag, and fuel economy. These problems are being inves-
tigated by Chrysler under a $6.5 million EPA contract awarded in
December 1972 for the development of an improved version of its sixth-
generation engine for powering an intermediate-size car. (Engine
design for a compact car will be examined in a recent extension to this
program. ) The NASA Lewis Research Center is playing an important
role in the effort in the areas of combustion, rotating machinery design,
and engine test.
5. While demonstrated fuel economy is low for gas turbine engine -
powered cars (about eight to nine miles per gallon for an intermediate-
size car over the Federal Emissions Test Driving Cycle), it is esti-
mated that fuel economy for improved engines will rise to about
12 miles per gallon and be competitive with that of current spark igni
tion engine-powered cars.
6. In the future, if ceramic turbines are developed, a substantial rise in
turbine operating temperature will be possible, leading to a significant
reduction in fuel consumption. For example, based on theoretical esti-
mates, a temperature increase at the turbine inlet from a current
H-2
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value of about 1800°F to a future value of 2500°F would result in about
a 20 percent improvement in fuel consumption. An additional benefit
might be reduced manufacturing cost.
RANKINE CYCLE ENGINES
1. Under the AAPS Program, the Rankine cycle engine is in an advanced
state of development, having progressed to the point where several
complete engine systems have been tested on engine dynamometers.
Some of these engine systems are water-base working fluid designs
(steam engine) that present the problem of fluid freezing at low tem-
peratures. Others rely on an organic-type working fluid as one means
of circumventing this problem.
2. Preliminary analyses of steady-state data from the AAPS Program
show that the emissions from this engine may meet the original 1976
Federal emission standards and that fuel economy can be competitive
with current emission-controlled spark ignition engines. A major
breakthrough in design of the engine condenser has resulted in much
smaller units, which has led to improved engine installation in current
automobile engine compartments.
3. The task of repeated demonstrations of low emissions in transient
operation (including cold start) with a fully automatic control system
has been initiated under the AAPS Program. Scientific Energy Systems
Corporation has been selected from several competing contractors to
provide EPA with a vehicle-installed steam engine system for testing.
This engine design utilizes a reciprocating piston expander for trans-
mitting power through a transmission to the rear wheels (in contrast
to other designs based on a turbine expander).
4. Future AAPS Program efforts are expected to concentrate on achieving
improvements in fuel economy and investigating low-cost production
techniques.
5. In addition to the EPA AAPS Program, the U.S. Department of Trans-
portation has investigated Rankine cycle power systems for buses; and
the State of California, under the Clean Car Project, has been testing
this type of engine for passenger car use. Other domestic and foreign
programs are being funded by private capital.
STIRLING ENGINES
1. The Stirling engine has the potential for demonstrating excellent fuel
economy, multifuel capability, very low noise and vibration, and
emissions low enough to meet the 1976 Federal standards.
H-3
-------
2. The most significant advancement in recent times is a major reduction
in engine volume, coupled with a major increase in power output per
pound of engine weight. This has led to the initiation of a Ford demon-
stration program that involves the installation of engines in Torino and
Pinto automobiles. N. V. Philips Laboratories at Eindhoven, Holland,
will furnish the engine for the Torino installation, and United Stirling
of Sweden will provide the Pinto engine.
3. Primary problem areas requiring further development work for reso-
lution are:
a. For efficient operation of this engine, the radiator in certain
designs can be about two and one-half times as large as that for
a comparable internal combustion engine. A reduction in
radiator size will be necessary to permit satisfactory installation
in current automobile engine compartments.
b. To contain the high-pressure hydrogen working fluid, the array
of heater tubes is now made of nickel-chrome alloys. These
tubes are expensive to fabricate and, therefore, production
processes must be improved or new designs must evolve to aid
in reducing engine manufacturing costs.
4. Other problems with the Stirling engine include inadequate piston seal
life, hydrogen diffusion through the cylinder walls and seals, and the
need for low-cost power control capable of providing smooth power
transitions. Engineering solutions are available, but they require
further development and demonstration.
DIESEL ENGINES
1. A number of manufacturers (e.g., Mercedes, Peugeot, Opel) are cur-
rently marketing a passenger car powered by a diesel engine.
2. Based on test data taken by EPA, research institutions, and
engineering/manufacturing firms, light-duty vehicles powered by diesel
engines have emissions below the original 1975 Federal emission
standards, but there are no indications as yet that this engine can ade-
quately meet the original 1976 Federal NOX standard. Additionally,
diesel odor and particulates (smoke) have long been recognized as very
undesirable exhaust emission products. Progress toward determina-
tion of the cause of the odor has been very slow, though its control has
been achieved in some engines through fuel injection system refinements.
3. The fuel economy of diesel-powered vehicles over the Federal Emis-
sions Test Driving Cycle is between 50 percent and 70 percent better
than that achieved by the average 1973 model year emission certifica-
tion vehicles tested at the same inertia weight. Currently, however,
H-4
-------
because installed diesel engine power is relatively low, vehicle
acceleration is significantly lower than that found with spark ignition
engine-powered cars. On an equivalent performance basis, the fuel
economy advantage of a higher-powered diesel vehicle would be expected
to be smaller than that for current lower-powered diesel vehicles.
4. Power output per pound of engine weight is considerably lower than that
presently offered in spark ignition engine-powered domestic automo-
biles. Methods for increasing the power output per pound of engine
weight include: a) means to increase power by turbocharging or super-
charging, and b) means to reduce engine weight and bulk by a reduction
in engine compression ratio and a shortened design life.
WANKEL AND OTHER ROTARY PISTON ENGINES
1. Rotary engine-powered car production levels have been rising, although
they represent a very small fraction of the automobile population.
Currently, these engines appear principally in Mazda vehicles manu-
factured by Toyo Kogyo. Production levels are expected to increase
with the limited introduction scheduled for the 1975 model year of the
Chevrolet Vega powered by the General Motors rotary engine.
2. For a given power rating, the rotary engine is lighter and smaller than
a reciprocating piston spark ignition engine. Reductions in engine size
and -weight can also lead to smaller and lighter vehicles as the result of
more compact packaging.
3. A manufacturing problem exists in attempts to increase production
rates. Because of complex machining operations required on both the
housing and rotor, production of this engine is currently limited to
about 25 per hour per line (compared to piston engine rates of over
100 per hour per line). This limitation appears to have a significant
impact on manufacturing costs.
4. Technological problems still exist for some engine designs. These
consist of high gas-seal leakage, high thermal stresses, poor low-
speed fuel economy, and high exhaust emissions without aftertreat-
ment. With aftertreatment (a thermal reactor), the Mazda rotary
engine has demonstrated the ability to meet the original 1975 Federal
emission standards.
STRATIFIED CHARGE ENGINES
1. Honda has a divided-chamber version (designated CVCC) of the strati-
fied charge engine in production. Without incorporation of additional
H-5
-------
emission control systems, its CVCC engine-powered Civic vehicle met
the original 1976 Federal emission standards for HC and CO, while
NOX was about twice the standard. Vega and Impala vehicles achieved
similar emission levels at low mileage when powered by General
Motors engines modified by Honda to the CVCC configuration.
2. Army M-151 vehicles equipped with Texaco's and Ford's experimental
open-chamber version of the stratified charge engine could meet the
original 1976 Federal standards with the use of an emission control
system incorporating oxidation catalysts and exhaust gas recirculation.
But these vehicles were unable to negotiate the high acceleration modes
of the Federal Emission Test Driving Cycle.
3. Fuel economy of the Honda Civic vehicle with the CVCC engine, as
measured over the Federal Emissions Test Driving Cycle, was about
ten percent lower than that of conventional Civic vehicles, and 16 per-
cent lower than that of equivalent-weight 1973 model year emission
certification vehicles. A Vega vehicle with this type of engine showed
fuel economy five to ten percent better than that of the standard Vega.
A CVCC powered Impala and a Texaco TCCS powered M-151 vehicle
showed equal or slightly better fuel economy compared to the respec-
tive unmodified vehicles.
4. Both open-chamber and divided-chamber engines have demonstrated a
lower sensitivity to fuel octane number than the basic spark ignition
engine.
5. A potential problem area for both open-chamber and divided-chamber
engines is high production cost. The open-chamber engine also
requires cylinder fuel injection. Additional potential problem areas
related to open-chamber engines include emission control system
durability and the current inability to achieve oxides of nitrogen emis-
sion levels that meet original 1976 Federal emission standards without
incurring a substantial loss in fuel economy.
H-6
-------
SUMMARY
-------
SUMMARY
The Clean Air Act Amendments of 1970 were the stimulus for
vigorous research efforts on the part of the automotive and related industries
to find the means to meet stringent exhaust emission standards for hydro-
carbon (HC), carbon monoxide (CO), and nitrogen oxides (NO ) required for
jC
1975 model year automobiles and beyond. Most of these efforts focused on
modifying the combustion process and/or treating exhaust gases of the con-
ventional spark ignition internal combustion engine. Some of the research
effort went into investigating the feasibility of low-emission alternative types
of engines.
Table S-l summarizes the applicable statutory emission
standards for gasoline -powered passenger cars as currently promulgated.*
As can be noted, the Federal standards apply nationwide except for those
instances where California has been granted a waiver to adopt more stringent
control of one or more emission species in a given time period.
Through model year 1974, these standards have been met by
the combination of engine modifications (combustion chamber redesign,
spark retardation, etc.) and the incorporation of exhaust gas recirculation
systems for control of nitrogen oxides to meet 1973 - 74 standards. In
regard to the 1975 California standards, it is generally agreed that the
required emission control system is exemplified by the following package of
components and engine modifications:
• Oxidation catalytic converters
• Air injection
•i*
''^Exhaust emission goals for engines under investigation in the AAPS Program
are set at one-half the 1977 Federal standards. The requirements are made
more severe to: (a) account for changes from prototype to mass-production
designs, (b) ensure that production engines, with expected statistical varia-
tions, can pass the Federal Certification Test, and (c) allow for expected
increases in emissions as vehicle mileage reaches the 50, 000-mile Federal
certification limits.
S-l
-------
Table S-l. A Summary of Passenger Car Emission Standards
(Gasoline Powered)
CO
1
ro
PP
Tei
Kmisi mn
t Procedure
Hydrocarbons
Carbon Monoxide
N it roperi Oxides
Evaporative
Emission Rate. Cramj/Milr
Unc ont rol led
Pre-I96fca
FTP CVS-C
10 17
77 125
4-6 4-6
40
1966 Cal
1968 Fed
1970
Cal Fed
I'm
C4] Fed
3. 4
35
NR
NR
2.2 I. I
23 23
NR NR
6 NR
Z.Z 2.2
23 23
4 NR
6
1972
Cal Fed
1973
Cal Fed
1974
Cal Fed
CVS-C
3.2 3. 4
39 39
3.2b NR
2
3.2 3.4
39 19
3.0 3.0
2
3.2 3.4
39 39
2.0 3.0
2
I97'ic
Cal Fed
197fcd
Nation
I977d
Nation
CVS-CH
0. 9 1. S
9. 0 15
2. 0 3.1
2
0.41
3 4
2.0
2
0 41
3. 4
0.4
2
FTP 7 d F d d
NR No requirement
CVS-C CVS cold start mass test
CVS-CH CVS cold/hot weighting mass test
Estimate only, for comparison purposes
Two hot cycles of FTP test
°Original 1975 standards were: HC= 0.41, CO= 3.4, NOx= 3.1
Original 1976 standards were the same as current 1977 standards
-------
• Exhaust gas recirculation
• Carburetor modifications
• Ignition system modifications.
(There is not complete agreement as to whether the catalytic converter will
be required to meet the less stringent Federal 1975 standards which apply to
the rest of the nation.) An example of emission control system elements for
control of both exhaust and evaporative emissions is given in Figure S-l for
General Motors' projected 1975 design.
Emission control systems to meet the 1976 standards are
projected to be similar to the 1975 systems, with improvements in catalytic
converters and associated components to meet the lower hydrocarbon and
carbon monoxide requirements.
Emission control systems which have been considered by the
automobile manufacturers to meet the stringent nitrogen oxide standard pre-
scribed for 1977 include all components of the 1974 system, plus:
• Reduction catalyst(s) installed upstream of the oxidation
catalyst(s)
• More sophisticated air injection systems
• Modified carburetion, ignition, and EGR systems.
To date, no manufacturer has reported acceptable durability for reduction
catalysts; however, further development testing is in progress.
The ability of the conventional internal combustion engine to
meet the original 1976 Federal exhaust emission standards has been the
subject of debate for some years. Even with the most elaborate emission
control schemes, prospects for meeting these standards are considered
marginal by most automakers. Further, the add-on emission control sys-
tems are expensive, they affect driveability of the car, their durability for
50, 000 miles is a serious problem, and they result in degraded fuel economy.
As an alternative, a number of other engines have been con-
sidered for the automobile power plant of the future. These engines, as
conceived, may not require add-on devices; in most cases, it is expected that
S-3
-------
-AIR INJECTION
PUMP
QUICK HEAT
MANIFOLD (EFE)
IMPROVED CARBURETION AND CHOKE
ALTITUDE AND TEMPERATURE
COMPENSATION
EXHAUST GAS
RECIRCULATION
MODIFIED SPARK
TIMING
CATALYTIC
CONVERTER
PCV VALVE
DOMED TANK
VAPOR SEPARATOR
-CARBON
CANISTER
ELECTRONIC
IGNITION
Figure S- 1 .
General Motors Projected 1975 Under-Floor Emission
Control System
-------
the basic combustion process can be adequately controlled to achieve the
expected low emission levels. In this regard, the most promising systems
rely on continuous combustion processes as opposed to the intermittent
combustion processes found in the spark ignition engine.
Evaluation of low-emission alternative engines to meet the
Federal requirements has been a continuing task since 1970, when its pro-
gram of research and technology development was launched by EPA to evalu-
ate the feasibility of various automotive power plants. The program is
called the Alternative Automotive Power Systems (AAPS) Program and is
administered by the Alternative Automotive Power Systems Division of EPA.
This program -was intended to supplement the work of industry, thereby
ensuring that all reasonable technical approaches to meeting the standards
were being pursued. A few of the engines under study were invented several
decades ago and have been used successfully as power generators for a num-
ber of years. However, until recent applications of modern-day technology,
these concepts -were considered too inadequate in weight, size, cost, and
efficiency for serious consideration as a mass-produced automobile power
plant.
Six types of power systems were initially included in the
program: gas turbine, Rankine cycle, heat engine/electric and heat engine/
flywheel hybrids, all-electric, and stratified charge engines. In July 1971 it
was decided to limit the number of candidate systems under development and
to concentrate all available funding and manpower resources on the three
most promising systems for the near term: the gas turbine, the Rankine cycle,
and stratified charge engines. The theoretical characteristics and anticipated
development schedules of these three systems were considered more favor-
able than the others. Consequently, development work is focusing on these
three, with major efforts going to the gas turbine and Rankine cycle engines
because of their greater need for new technological advancements. The pro-
gram is schedule to be completed in 1975 with demonstrations, tests, and
evaluations made of Rankine cycle, gas turbine, and stratified charge engines
in automobiles.
S-5
-------
This volume of the report to the Administrator is intended to
summarize the status of engine development work in the AAPS Program
and of other similar hardware development programs on alternative engines
being funded by private industry as of the end of 1973 . The engines under
review are the: gas turbine, Rankine cycle, Stirling, diesel, Wankel (and
other rotary piston designs), and stratified charge. Of these, gas turbine,
Rankine cycle, and Stirling cycle engines operate with continuous combustion
processes; whereas, the diesel, Wankel, and stratified charge (and the
current spark-ignition) engines operate with intermittent combustion
processes. (Continuous combustion processes offer the engine designer
more flexibility in tailoring combustion for low emissions.) A discussion
of current progress in development of these engines is summarized in the
next section.
S. 1 PROGRESS IN ENGINE DEVELOPMENT
S. 1. 1 Gas Turbine Engine
S. 1. 1 . 1 General Description
In its simplest form, the gas turbine engine consists of a
compressor, a combustion chamber, and a turbine (Figure S-2). Air is
taken in by the compressor at atmospheric pressure, compressed to a higher
pressure, then delivered to the combustion chamber where fuel is injected
and burned at essentially constant pressure. The resulting high-temperature,
high-pressure gas is expanded in the turbine and then exhausted to the atmo-
sphere. Part of the turbine-shaft work is used to drive the compressor; the
remainder is the output work.
Several modifications of this simple cycle have been employed
in the development of gas turbine engines for vehiclar use. Most current
engines for on-highway vehicles have a free power turbine (Figure S-3); i.e.,
two mechanically independent turbine stages are used. One drives the com-
pressor and the other drives the output shaft. This allows a wide range in
output-shaft speed, which is not possible with a single-shaft turbine engine.
S-6
-------
Fuel in
Air intake
s
p
Combustor
/
\
Outp
shaft
Compressor
Turbine
Figure S-2. Schematic Diagram of Simple Gas Turbine Engine
Fuel in
Air intake
Exhaust
Combustor
Compressor
Output shaft
Compressor
turbine Power
turbine
Gas-generator section
Figure S-3. Schematic Diagram of Free Turbine Engine
S-7
-------
To improve efficiency, most free-turbine engines used a regenerative
or recuperative cycle, incorporating a heat exchanger that delivers heat from
the exhaust gas to the compressed air upstream of the combustor. The heat
exchanger is called a recuperator, if it has a fixed surface, or a regenerator
(Figure S-4) if it is a rotary type. The heat exchanger decreases the exhaust
temperature but increases the engine weight and volume.
S. 1. 1. 2
Historical Development
The fundamental principles of the gas turbine engine were
well known by the end of the eighteenth century; however, turbomachinery
operating successfully on the gas turbine thermodynamic cycle, called the
Brayton cycle, is a fairly recent development. The first really successful
gas turbine power plant was the aircraft turbojet developed from intensive
work begun in the 1930s.
Development of gas turbine engines for automotive uses was
begun immediately after World War II with Chrysler in this country and
Exhaust
Compressor
Compressor
turbine
Output
3 shaft
Power
turbine
Gas generator section
Figure S-4. Schematic Diagram of Regenerative
Free Turbine Engine
S-8
-------
Rover in Great Britain. The first Rover turbine-power automobile ran in
1949. General Motors began automotive gas turbine work around 1948;
Ford about 1950. All of these developers selected the free turbine configu-
ration early and moved quickly to the regenerative configuration. In 1964,
Chrysler fabricated and tested a demonstration fleet of 50 gas turbine-
powered automobiles. More recently, General Motors and Ford's principal
gas turbine engine developments have focused on engines for bus and truck
applications instead of for passenger cars. To date, no such units are
commercially available; but in some cases prototype engine-vehicle testing
is continuing.
In 1971, the AAPS gas turbine program was initiated to
attempt to solve the technical problems that had plagued past attempts at
developing a production gas turbine for the passenger car. The basic ele-
ments of the program are shown in Figure S-5. In this type of program,
automotive gas turbine engines are initially used as test beds to establish
baseline performance and emission levels. Then they are used to verify
component improvements resulting from the technology programs sponsored
in" the areas of low-emission combustion, manufacturing cost studies,
improved part-load fuel economy studies, materials development, and
transmission and controls studies. An engine improvement phase is included
that allows redesign of the gas turbine according to the technology advance-
ments made. The development team selected to implement the program is
shown in Figure S-6.
Estimated high production costs continue to be a major factor
inhibiting the implementation of the automotive gas turbine engine. Other
problems involve the need for improvements in noise level, durability,
acceleration lag, and fuel economy. These problems are being investigated
by Chrysler under a $6. 5 million EPA contract awarded in December 1972
for the development of an improved version of its sixth-gene ration engine
for powering an intermediate-size car. (Engine design for a compact car
will be examined in a recent extension to this program.)
S-9
-------
C/5
I
PROGRAM PHASE
EPA TECHNOLOGY PROGRAMS
COMBUSTION TECHNOLOGY
MANUFACTURING (low cost)
COMPONENT DEVELOPMENT
NASA TECHNOLOGY PROGRAMS
COMBUSTION TECHNOLOGY
MANUFACTURING (low cost)
COMPONENT DEVELOPMENT
NASA POWER SYSTEM IMPROVEMENT
PROGRAM
COMPONENT TESTING
AERODYNAMIC IMPROVEMENT
ENGINE DYNAMOMETER TESTING
CHRYSLER BASE LINE ENGINE PROGRAM
COMPONENT IMPROVEMENT
ENGINE DYNAMOMETER TESTING
VEHICLE TESTING
UPGRADED ENGINE DESIGN AND TEST
EPA VEHICLE EMISSION TESTING
AND EVALUATION
CALENDAR YEAR
1971
I
Fl
i
1972
i
i
FINAL
vlAL DEN
i
1973
1974
1
i
i
L
1
CT
cc
\ M \ X
1975
•v
/
>k
s
i
| \ XI
iVALUA
XI
CO
j
. roi
VER
X
TION REPORT-
L
Ml
IF
Flh
UP
IF
1976
^ONENT
LOW
ML
ONENT
ICATION
*
t
rfONSTRATION TEST— J
i i
Figure S-5. Gas Turbine Power System AAPS Program
-------
AAPS
BRAYTON POWER SYSTEMS DEVELOPMENT TEAM
r
[NASA LEWIS TECHNOLOGY PROGRAMS
COMBUSTOR HEAT EXCHANGER MANUFACTURING
(low cost)
CATALYTIC [-OWENS ILLINOIS [-AIRESEARCH
(inhouse) L-CORNING I-PRATT & WHITNEY
SOLAR
GENERAL ELECTRIC
AEROJET LIQUID
ROCKET CO.
I
[ EPA
r
COMBUSTOR
[-WILLIAMS RESEARCH
-SOLAR
-AIRESEARCH
-PRATT & WHITNEY
-MECHANICAL
TECHNOLOGY, INC.
-NORTHERN RESEARCH
-GENERAL ELECTRIC
L AEROJET LIQUID
ROCKET CO.
SYSTEM IMPROVEMENT
TECHNOLOGY PROGRAMS |
HEAT EXCHANGER MANUFACTURING STUDIES
(low cost)
-OWENS ILLINOIS I- AIRESEARCH
-CORNING 1- PRATT & WHITNEY
' 1
ECONOMIC ADVANCED TURBINE
l-WILLIAMS DESIGN
RESEARCH [-GENERAL ELECTRIC
1- AIRESEARCH
1- PR ATT & WHITNEY
_L
NASA LEWIS
'
t
SYSTEM
IMPROVEMENT
COMPONENT
TESTING
POWER SYSTEM
TESTING
AERODYNAMIC
IMPROVEMENT
j-TURBINE
[-COMPRESSOR
LFLOW PASSAGES
SYSTEM
IMPROVEMENT
-COMPONENT
TESTING
-POWER SYSTEM
TESTING
u VEHICLE
TESTING
_L
CHRYSLER
COMPONENT
IMPROVEMENT
-CONTROLS
-HEAT EXCHANGER
-TRANSMISSION
-INHOUSE COMBUSTOR
-GOVERNMENT FURNISHED
EQUIPMENT (from NASA
Technology Programs)
-NOZZLE ACTUATOR
-FREE ROTOR
GAS TURBINE
UPGRADING
Figure S-6. EPA/AAPS Brayton Power Systems Development Team
-------
Estimates and measurements of current gas turbine fuel
economy capability are given in Table S-2. While demonstrated fuel
economy is low (about eight to nine miles per gallon for an intermediate-size
car over the Federal Emissions Test Driving Cycle), it is estimated that fuel
economy for improved engines will rise to about 12 miles per gallon and be
competitive with that of current spark ignition engine-powered cars. In the
far-term period (1985 to 2000), if ceramic turbines are developed, a sub-
stantial rise in turbine operating temperature will be possible, leading to a
significant reduction of fuel consumption. For example, based on theoretical
estimates, a temperature increase at the turbine inlet from a current value
of about 1800°F to a future value of 2500°F would result in about a 20 percent
improvement in fuel economy.
S. 1. 1.2 Current and Projected Status
Principal early automotive-turbine problems were high cost,
noise, poor fuel economy, durability, and acceleration lag. Considerable
improvements have been made in each of these areas over the years. It is
expected that the next generation of gas turbine engines will be somewhat
further improved in these areas, particularly in fuel economy. The recent
awareness of domestic energy shortages has placed additional emphasis on
improving fuel economy in the gas turbine program sponsored by the Alter-
native Automotive Power System Division of EPA. In the far term, if
ceramic turbines are developed, a substantial rise in turbine inlet tempera-
ture would be possible, leading to a significant reduction in fuel consumption.
With regard to exhaust emissions, those of automotive-type
gas turbine engines have been typified until recently as characteristically
low in unburned hydrocarbon/carbon monoxide, but with nitrogen oxide levels
too high to permit attainment of the original 1976 Federal emission standard
S-12
-------
Table S-2. Gas Turbine Fuel Economy Data Summary
Engine
United Aircraft
Research Laboratoriesa
RGSS-6
(Regenerative
Single Shaft)
RCSS-8
(Recuperative
Single Shaft)
SS-10
(Simple Cycle)
AiResearcha
Free-turbine
Regenerated
Chrysler
Free-turbine
Regenerated
(sixth- generation
engine)
Williams Research
WR26 Engine
Free-turbine
Regenerated
Rated
Power
Level
at 59°F
150
150
150
175
150
80
Am-
bient
Temp.,
°F
59
59
59
59
Fuel Economy, mi/gal
FETDC
12. 16
11.20
7 51
12.5
8.5b
5.1, 8.1b
20
mph
15.06
13.73
8.86
16. 1
30
mph
18.02
16.59
11.21
18.2
40
mph
18.65
17.33
12.51
18.2
18
—
50
mph
17.96
16.73
12.63
17.2
60
mph
16.59
15.42
12.22
15.5
_ _
70
mph
14.70
13.79
11.47
13.7
—
Vehicle
or Test
Weight, Ib
4000
3950
3700
4000
Inter-
mediate
size car
AMC
Hornet
(compact)
Estimates based on computer simulation. mi/ gal - miles per gallon
bBased on tests over Federal Emissions Test Driving Cycle (FETDC). mph = mlles Per hour
-------
of 0.4 grams per mile. For example, Chrysler's sixth-generation engine
installed in an intermediate-size vehicle has nitrogen oxide emissions of
about 2. 2 grams per mile over the Federal Emissions Test Driving Cycle.
However, within the AAPS Program, Solar, a Division of International
Harvester; has recently developed a new gas turbine combustion system that,
based on steady-state test data, is calculated to be capable of meeting the
original 1976 Federal emission standards and is close to meeting the AAPS
Program goals (one-half the level of the original standards). Additional
development is required to perfect the combustor control system and to
validate meeting program goals with transient operation of the engine over
the Federal Emissions Test Driving Cycle.
Reflecting similar recent advancements a General Motors-
sponsored program resulted in an experimental 225-horsepower gas turbine
engine installed in a car meeting the original 1976 Federal emission stan-
dards. Complete details concerning engine design and operation are not
available. The available gas turbine data are summarized in Table S-3.
Estimates of the manufacturing costs of gas turbine engines
in mass production differ widely. A recent cost analysis conducted under
the AAPS Program indicated somewhat higher initial turbine engine cost, as
compared to a piston engine, but a potential for a slightly lower cost with
further development. While encouraging, it is recognized that actual costs
are typically higher than projected costs by the time the hardware is devel-
oped, tested, and produced. The study nevertheless provided valuable
insight into processes that must be developed and components whose designs
and materials must be simplified and improved in order to have a low-cost
engine. Several cost-reduction programs are under way for processing cur-
rent metal components. The successful development of ceramics for sta-
tionary hot parts and, in the longer term, for rotating parts would lower
the production cost of the gas turbine significantly and conserve the coun-
try's scarce resources of alloying materials.
S-14
-------
Table S-3. Gas Turbine Exhaust Emission Data Over the
Federal Emissions Test Driving Cycle
Engine /Manufacturer
a. Conventional Combustors
United Aircraft Research
Laboratories RGSS-6
United Aircraft Research
Laboratories SSS-10
Chrysler sixth-generation
engine
Williams Research/
WR26/AMC Hornet
Williams Research/
131Q/Volkswagen
b. Advanced Combustors
Solar
General Motors 225 HP
Regenerative Engine
AiResearch
United Aircraft of Canada
c. Clean Air Act Requirements
Emissions (gm/mi)
CO
0. 53
1.86
3.99
7.43
6.92
4.5
3.34
2.4
1.6
3. 64
3.4
HC
0. 15
0. 31
-0.26
0. 62
0.72
0. 34
0. 11
0. 015
0. 67
0.49
0.41
NO
X
2.72
1. 03
2.21
2.8
2. 5
1.81
0. 34
0. 315
0. 44
0.52
0. 4
Remarks
Calculated values over simulated
Federal Driving Cycle, based on
emission data from GM engine
GT-309.
Calculated values over simulated
Federal Driving Cycle, based on
emission data from GM engine
T-56.
1975 Federal Test Procedure
emissions corrected (background
level subtracted from measured
values).
Cold start /1975 Federal Test
Hot start (Procedure
1972 Federal Test Procedure.
Calculated values for simulated
Federal Driving Cycle, based on
advanced combustor data.
Experimental test bed engine,
chassis dynamometer test with
5,000-lbcar, 1975 Federal Test
Procedure.
Calculated values for simulated
Federal Driving Cycle, for recu-
perated engine with 10 percent
bypass .
Calculated values over simulated
Federal Driving Cycle based on
advanced combustor data.
Original 1976 Federal standards
S-15
-------
S.I.2 Rankine Cycle Engine
S.I.2.1 General Description
The Rankine cycle engine is an external combustion engine in
which high-pressure steam or some other working fluid is expanded in a
reciprocating (piston-type) or turbine-type device to produce work. Compo-
nents of a typical Rankine cycle engine are shown in schematic form
(Figure S-7). The feed pump draws liquid at a low pressure from the reser-
voir and forces it under pressure into the vapor generator, where it is con-
verted to superheated vapor by heat from the combustion gases. The hot,
high-pressure vapor then flows into the expander where it gives up energy as
it expands to a lower pressure and temperature. Exhaust vapor from the
expander enters the condenser, where it is converted back to a liquid by
rejecting heat to the surroundings. The resulting low-pressure liquid is then
returned to the reservoir or sump to complete the cycle. The major com-
ponents shown (feed pump, vapor generator, expander, condenser) are
required for any Rankine cycle system. Specific systems will show varia-
tions in such features as type of drive for the auxiliaries, additional heat
transfer paths from one part of the cycle to another, and the type and location
of liquid reservoir and boost pump.
S. 1. 2. 2 Historical Development
Rankine cycle steam cars held a prominent place in the early
days of the automobile. Steam cars were sold commercially in the late
1800s and early 1900s with familiar names such as Stanley Steamer, Loco-
mobile, White, and Doble. These early steam cars had a simple pressurized
boiler containing a large quantity of water; which required about a 30-minute
cold startup and posed a continual safety problem. They were noncondensing,
dumping exhaust steam overboard. Consequently, frequent stops were
required for makeup water. In spite of these disadvantages, some of these
cars, such as the Stanley Steamer, were quiet running and performed quite
satisfactorily. Various technical advances were achieved through the years,
S-16
-------
CO
I
ELECTRIC
MOTOR
ATOMIZING
AIR
PUMP
•CXI
THROTTLE
VAPOR
GENERATOR
HYDRAULIC
AIR PUMPS
INTAKE
AUXILIARY!
POWER l
ACCUMULATOR
HYDRAULIC
MOTOR
EXHAUSTn
HYDRAULIC
MOTOR
J
I
FEED
PUMP
TRANSMISSION,
DRIVE WHEELS
FAN
HYDRAULIC
MOTOR
SUMP
SUBCOOLER
<\ I
AIR-
COOLED
CONDENSER
MAKEUP
PUMP
Figure S-7. Rankine Cycle System Schematic
-------
primarily the monotube boiler by Doble, which effectively solved the safely
and startup problem. Condensing radiators were in use from about 1915 on,
resulting in true close-Rankine-cycle operation. By this time, however,
the internal combustion engine entered into its era of low-cost mass produc-
tion and, in this respect, the steam car did not compete, disappearing from
the automotive scene in the very early 1930s. The reasons for the lack of
similar cost-reduction development for the early steam cars may have been
varied, but they do not appear to have been primarily technical.
Interest in the steam car occurred sporadically during the
fifties and early sixties; but it was not until the latter sixties, with the
impact of impending air pollution control legislation becoming apparent, that
the Rankine cycle automobile system began to receive serious scientific
attention. The impetus behind this renewed interest lay in the clean-burning
characteristics of the external combustor of this system, which requires
neither post-engine exhaust cleanup devices nor compromises with an effi-
cient burning process.
A number of engineering efforts were undertaken at about this
time by widely diverse organizations. This development received impetus
and cohesion with the initiation in 1970 of the EPA Alternative Automotive
Power System Program. Figure S-8 depicts the AAPS Rankine program
schedule, and Figure S-9 delineates the development team selected to imple-
ment the program. This EPA effort was structured to include problem
solving and component development activities. Component performance
improvements were incorporated into preprototype engine systems. The
final phase of development on this program is the prototype system and
vehicle demonstration phase.
S. 1. 2. 3 Current and Projected Status
Table S-4 summarizes the more significant worldwide Rankine
cycle engine efforts both completed and in progress. Under the AAPS Pro-
gram, this engine is in an advanced state of development, having progressed
S-18
-------
Cfl
I
h-^
xO
PROGRAM PHASE
TECHNOLOGY PROGRAMS
PREPROTOTYPE DEVELOPMENT:
EMISSIONS
PREPROTOTYPE DEVELOPMENT:
EMISSIONS, FUEL ECONOMY
PROTOTYPE DEVELOPMENT:
EMISSIONS, FUEL ECONOMY, NOISE
VEHICLE DEMONSTRATION:
CALENDAR YEAR
1971
1972
1
1 SYS
FOR
1973
1974
| | 4 SYSTEMS
TEM SEL
3ROTOTY
ECTED A
PE t
Fl
4 OR LE
SYSTEM
1975
:ss
s
1
>IAL EVA
REPO
1 A
LUATION
RT
Figure S-8. Rankine System Prototype Development Program Schedule
-------
c/i
i
(SJ
O
ENVIRONMENTAL PROTECTION
AGENCY
INASA LERC L» «-r — r n-ir — i _ •_ _
|
1 1 1
CONTRACT TECHNICAL R&D
MONITORING SUPPORT PROGRAMS
SYSTEM CO
I
(LOGY PROGRAMS ]
1
1
[COMBUSTOR | | CONDENSER |
\ITR
-SOLAR
GEO SCIENCE
LTD
1
MODELING
STUDIES
- A^sl ARCH L GENERAL
AIRESEARCH ELECTRIC
-UNIV. OF MICHIGAN
1 1
1
-KAAVD | LUBRICATION | | FLUIDS) | FEEDPUMP
-BATTELLE 1 1
L ELECTRIC LMONSANTO
ACTORS
-LEAR
• CHANDLER
EVANS
1
I
1
1
I WATER-TURBINE I | WATER-RECIPROCATORj | ORGANIC WITH RECIPROCATOR | j ORGANIC WITH TURBINE |
I . I _. I
I LEAR MOTORS j JSTEAM ENGINE SYSTEMS*! | THERMO ELECTRON CORP \
\
GENERAL MOTORS
-RICARDO
-CHRYSLER
-ESSO
-BENDIX
L
FORD
| AEROJET!
LGENERAL
MOTORS
*(now Scientific Energy Systems Corp)
Figure S-9. EPA/AAPS Rankine System Development Team
-------
Table S-4. Summary of Rankine Cycle Engine/Vehicle Development Programs
Developer /Contractor
EPA AAPS Program
Aerojet Liquid Rocket Co.
Lear Motors Corp.
Scientific Energy Systems
Thermo Electron Corp.
DOT/California Steam Bus Program
Brobeck &t Associates
Lear Motors Corp.
Steam Power Systems Co.
DOT/Dallas Program
Sundstrand Corp.
California Clean Car Project
Aerojet Liquid Rocket Co.
Steam Power Systems Co.
Carter Enterprises, Inc.
Kinetics, Inc.
Williams Engine Co.
Foreign Programs
Pritchard Steam Power Proprietary, Ltd.
SAAB-SCANIA
Politecnico Milano
Vehicle
6 -Passenger cara
5 -Passenger cara
6 -Passenger cara
6 -Passenger cara
51-Passenger bus
51-Passenger bus
51-Passenger bus
25-Passenger bus
4 -Passenger cara
4-Passengcr cara
4-Passenger car
(Volkswagen)
4-Passenger car
4 -Passenger car
6- Passenger ca r
Passenger car3
Bus*
Rated (or Estimated)
Horsepower
ISO, Expander gross
123, Expander gross
158, Expander gross
1 38, Transmission in
146, Expander gross
240, Gross bhp
200, Net bhp
240. Gross bhp
180, Net bhp
275, Gross bhp
224. Net bhp
120, Expander ijross
90, Gear box out
65, Expander gross
70
160
100
aSimulated or installation design target. bhp - brake horsepower DOT -
Type of Expander
and Rated Speed
in rpm
Turbine; 32,000
Turbine; 65,000
Piston; 2, 500
Piston; 1, 800
Piston; 2, 100
Turbine; 65,000
Piston; 1,850
Turbine; 35,000
Piston; 2,400
Piston; 5,000
Gerotor type
Piston
Piston
Piston
Turbine
Working Fluid
AEF-78 (Proprietary)
Water
Water
Fluorinol - 85
Water
Water
Water
CP-25; a single chemical
compound comprised of
methyl benzene, toluene,
and toluall.
Water
Water
Water
Refrigerant 113
Water
Water
Water
Organic
D t f T t t"
P P
Status of Program
Preprototype engine devel-
opment program completed
in December 1973. One
contractor (SES) selected to
proceed with prototype eng-
ine development, followed
by installation in vehicle.
Program completed Sept-
ember 1972. Consisted
of demonstration use in
public service and vehicle
emissions testing.
Program completed.
transit system and vehicle
emissions testing.
Testing and vehicle instal-
lation in progress. Target
date for vehicle delivery
is May 15, 1974.
Vehicle first driven March
ments being made.
No published emissions or
fuel economy data.
Component development.
Recently initiated.
en
-------
to the point where several complete engine systems have been tested on
engine dynamometers. (Other programs listed in Table S-4 have produced
demonstration vehicles that have shown system feasibility.) Some of these
engine systems are water-base working fluid designs (steam engine) that
present the problem of fluid freezing at low temperatures. Others rely on
an organic-type working fluid as one means of circumventing this problem.
A major breakthrough in design of the engine condenser has resulted in much
smaller units, which has led to improved engine installation in current
automobile engine compartments.
Preliminary analyses of steady-state data from the AAPS
Program show that exhaust emissions from the Rankine cycle engine may
meet the original 1976 Federal standards and that fuel economy can be com-
petitive with current emission-controlled spark ignition engines. Table S-5
summarizes available emission estimates for four classes of Rankine cycle
engines. Figure S-10 compares both current experimental and projected
Rankine cycle fuel economy characteristics with values obtained with two
prototype conventionally powered 1975 model year automobiles.
The task of repeated demonstrations of low emissions in tran-
sient operation (including cold start) with a fully automatic control system
has been initiated under the AAPS Program. Scientific Energy Systems
Corporation (formerly Steam Engine Systems) has been selected from
several competing contractors to provide EPA with a vehicle-installed steam
engine system for testing. This engine design utilizes a reciprocating piston
expander for transmitting power through a transmission to the rear wheels
(in contrast to other designs based on a turbine expander).
Future AAPS Program efforts are expected to concentrate on
achieving improvements in fuel economy and investigating low-cost produc-
tion techniques.
S-22
-------
Table S-5. Preliminary Rankine Cycle Emission Results Over the
Federal Emissions Test Driving Cycle
Engine Class
Water Reciprocating
Water Turbine
Organic Reciprocating
Organic Turbine
Emissions, gm/mi
HC
0.09
0.21
0.02
0. 16
CO
0.50
0.70
0. 19
1.01
NO
X
0.20
0.38
0.25
0.13
aThese values are based upon steady- state measurements
and are analytically converted to the 1972 Federal Test
Procedure.
gm/mi = grams per mile
-
20
,9 _
18
I '7
I"
I '5
E 14
>
8 '2
u
g 11
2 10
DATA CORRECTED TO SAME 4600-lb INERTIA WEIGHT AND TEST CONDITION
RANKINE CYCLE, PROJECTED (DEC 73)
1975 PROTOTYPE VEHICLES:
(spark ignition engine)
RANKINE CYCLE.
EXPERIMENTAL
ENGINE DYNAMOMETER -
(SEPT 731
20
30 40
SPEED, miles per hour
50
60
70
Figure S-10. Comparative Fuel Economy for
Rankine Cycle Engine
S-23
-------
S.I.3 Stirling Engines
S.I.3.1 General Description
The Stirling engine is an external-combustion, closed-cycle,
piston-type power plant that uses a gaseous internal working fluid, usually
hydrogen or helium (see Figures S-il and S-12). Cyclical heating and cool-
ing varies the pressure of the fluid within a closed volume, the pressure
variations being transmitted to a piston, thereby developing output power.
Heat transfer processes are accomplished at high efficiency by alternately
and regeneratively moving the working fluid between heater and cooler sec-
tions. Thus, the engine contains at a minimum one expansion volume, one
compression volume, one regenerator, one heater, and one cooler.
The ideal efficiency of the Stirling engine is 100 percent of
the Carnot efficiency obtained between peak and minimum fluid temperatures.
Various losses reduce actual efficiencies to about one-half of the ideal value.
Nevertheless, the Stirling engine is one of the most efficient prime movers
known.
S. 1. 3 . 2 Historical Development
Not a recent development, the Stirling-cycle engine was
patented in 1816 by Robert Stirling, a Scottish clergyman. In 1938,
N. V. Philips Laboratories at Eindhoven, Holland, undertook the develop-
ment of a modern version of this engine. Its research resulted in the devel-
opment of the rhombic drive displacer engine in 1953 and the roll-sock seal
in I960. This configuration had both the power piston and displacer piston
in a single cylinder. The rhombic drive made straight line rod movement
possible and opened the way to high working pressures and high power.
Displacer-type Stirling engines are operating in buses and boats in Europe.
In 1970, Philips initiated the most significant advancement yet
by achieving a major reduction in engine volume and complexity, coupled with
a major increase in power output per pound of engine weight. This occurred
through the design of and successful operation of a double-acting Stirling
engine with swashplate drive mechanism. This engine consists of
S-24
-------
HEATER
REGENERATOR
COOLER
0 PHASE
ANGLE
DISPLACER
PISTON
POWER
PISTON
SYNCHRONIZER
Figure S-ll. Schematic, Basic Stirling Displacer Engine
DISPLACER
PISTON
POWER
PISTON
Figure S-12. Schematic, Rhombic Drive Displacer Engine
S-25
-------
four separate, interconnected cylinders in which the piston motion is phased
at 90-degree intervals by the action of the swashplate. Each piston serves
alternately as a power piston and as a displacer piston for the adjacent
cylinder.
S. 1.3.3 Current and Projected Status
Philips is now developing an automotive engine in which the four
cylinders are equally spaced around a circle with a rotating wobble (or swash)
plate converting the reciprocating movement of the pistons into a rotating mo-
tion (Figure S-13). A 1971 agreement with Ford Motor Company is currently
scheduled to lead to the demonstration of such an engine in a Torino car in 1975.
United Stirling in Sweden is working on a V-4 configuration with conventional
connecting rods and crankshaft that Ford will install in a Pinto car. M.A.N/MWM
in Germany is also working on improvements and cost reduction designs.
Incentives for continuing work on this engine are excellent
fuel economy, multifuel capability, very low noise and vibration, size and
weight comparable to current engines, and emissions potentially low enough
to meet the original 1976 federal standards. Table S-6 compares projected
Stirling engine fuel economy characteristics with those estimated for a spark
ignition engine-powered automobile equipped with 1976-type emission con-
trol systems. Table S-7 compares Stirling engine exhaust emissions (simu-
lated) with the original 1976 Federal emission standards. The projected
emission levels fall well below the standards.
Primary problem areas with the Stirling engine requiring fur-
ther development work for resolution are the radiator size and the expense
of heater tubes. For efficient operation of this engine, the radiator in cer-
tain designs can be about two and one-half times as large as that for a com-
parable internal combustion engine. A reduction in radiator size will be
necessary to permit satisfactory installation in current automobile engine
compartments. To contain the high-pressure hydrogen working fluid, the
S-26
-------
co
cv
-J
IGNITOR
ATOMIZER
REGENERATOR
COOLER
TUBES
BURNER
ROTARY •
PREHEATEH
DOUBLE-ACTING
PISTON
OIL
PUMP
HEATER
TUBES
WATER
INLET
WATER
OUTLET
OUTPUT
SHAFT
GUIDE PISTON
AND SLIDERS
ROLL SOCK
PISTON ROD
Figure S-13. Philips Stirling Double-Acting Engine with Swash-Plate Drive
-------
Table S-6. Stirling and Spark Ignition Engine Fuel
Economy Comparisons
Driving Cycle
Consumer Average
Suburban Driving
City Traffic
Supe rhi ghway
Fuel Economy, mi/gal
1976 Ford V8
10.7
13.3
8.2
13.0
Stirling- Torino
(Projected)
14.7
17.9
11.4
16.6
Gain,
%
37
35
39
28
mi/gal = miles per gallon
Table S-7. Constant Volume Sampling Test
Simulation Emissions
Comparison
Philips Engine
Original 1976 Federal Standard
Emissions, gm/mi
HC
0. 10
0.41
CO
0.31
3.40
NO
X
0. 175
0.4
gm/mi = grams per mile
S-28
-------
array of heater tubes is now made of nickel-chrome alloys. These tubes are
expensive to fabricate and, therefore, production processes must be improved
or new designs must evolve to aid in reducing engine manufacturing costs.
Other problems with the Stirling engine include inadequate
piston seal life, hydrogen diffusion through the cylinder walls and seals, and
the need for low-cost power control capable of providing smooth power tran-
sitions. Engineering solutions are available, but they require further
development and demonstration.
S.I.4 Diesel Engine
S.I.4.1 General Description
The diesel or compression ignition engine is a reciprocating
engine which is, in many respects, very similar to the Otto cycle spark
ignition engine. In the diesel, air alone is compressed in the cylinder, and
the fuel is then injected into the heated air towards the end of the compres-
sion stroke. The temperature developed during compression is sufficiently
high to ignite the fuel immediately after injection. Thus, spark plugs, dis-
tributor, and carburetor are eliminated in the diesel; but a high-pressure
fuel-injection system is required. Since compression-ignition demands
high-pressure ratios, the current diesel engines are heavier, more bulky,
and costlier than equivalent spark ignition engines.
The principal advantages of diesels relative to spark ignition
engines include their ruggedness and their higher thermal efficiency, which
results in lower fuel consumption, particularly at part-load operating
conditions.
Many different engine configurations have been built around
the basic thermodynamic diesel cycle, and a number of these are currently
being marketed by many companies throughout the world. The various designs
can be grouped into two categories -- open chamber and divided chamber --
and each of these can then be subdivided into different engine classes, depend-
ing upon specific features related to the design of the combustion chamber
and the induction system. Open-chamber designs are.favored by most
manufacturers of heavy-duty diesels because of the better fuel economy
S-29
-------
obtained with this engine type. Conversely, divided-chamber designs
(prechambers and swirl chambers) are preferred by the manufacturers of
light-duty engines because of their higher speed capability.
S.I.4.2 Historical Development
The diesel engine was invented by Rudolf Diesel in 1892 and
was originally intended to burn coal dust. In his early design, Diesel pro-
posed to coordinate the rate of fuel injection with the movement of the pistons
in such a manner that the heat of combustion would be liberated at constant
temperature, thus approaching the thermal efficiency of the Carnot cycle.
However, this approach proved to be unsuccessful and the cycle was then
modified to achieve near-constant pressure combustion. While this ideal
process is still being utilized in large, low-speed diesels, the high-speed
diesels used in light-duty applications employ a process that lies between
the ideal cycle and the Otto cycle.
The high compression ratio required to achieve fuel ignition,
and the resulting bulk, precluded the application of diesels in early road
vehicles. The first diesel-powered truck was built in 1907 by Saurer in
Switzerland, in cooperation with Diesel; and in 1911 the first fully loaded
diesel truck crossed the United States coast to coast.
Daimler-Benz of Germany started series production of its
Mercedes Benz 260D diesel-powered automobile in 1936. Since resumption
of production in 1949, Daimler-Benz has sold more than one million diesel
automobiles. A number of other companies have since entered the light-duty
diesel engine field, including Peugeot of France, Perkins of England,
Daihatsu of Japan, and Opel of Germany.
S.I.4.3 Current and Projected Status
Because of their favorable fuel economy characteristics,
diesel engines are widely used in Europe and Japan as power plants for taxi-
cabs, passenger cars, and other light-duty vehicle applications. However,
the power output of these engines is considerably lower than that required
S-30
-------
for the heavier American vehicles, and their duty cycles are generally
quite different.
The Mercedes Benz 220D diesel-powered automobile has been
marketed in the United States since 1956. Between 1956 and 1972, Daimler-
Benz sales of diesel cars in the United States amounted to about 53, 000 units,
with more than 6, 000 sold in 1972 alone. In 1971, Daimler-Benz produced
about 284, 000 automobiles and 99, 000 of these were diesel-powered 220Ds.
Peugeot of France has introduced its diesel-powered 504-
passenger car in the United States in the spring of 1974. Since production
began in 1959, Peugeot has sold more than 480, 000 diesel-powered, light-
duty vehicles. In 1972, Peugeot's diesel sales amounted to about
80, 000 units out of a total production of 671, 000 vehicles.
Based on test data taken by EPA, research institutions, and
engineering/manufacturing firms, light-duty vehicles powered by diesel
engines have emissions below the original 1975 Federal emission standards
(Table S-8), but there are no indications as yet that this engine can ade-
quately meet the original 1976 Federal oxides of nitrogen standard. Addi-
tionally, diesel odor and particulates (smoke) have long been recognized as
very undesirable exhaust emission products. Progress toward determination
of the cause of the odor has been very slow, though its control has been
achieved in some engines through fuel injection system refinements.
The fuel economy of diesel-powered vehicles over the Federal
Emissions Test Driving Cycle is between 50 percent and 70 percent better
than that achieved by the average of 1973 model year emission certification
vehicles tested at the same inertia weight. Table S-9 lists some specific
comparative data for diesel- and gasoline-powered vehicles. Currently,
however, because installed diesel engine power is relatively low, vehicle
acceleration is significantly lower than that found in cars powered with spark
ignition engines. On an equivalent performance basis, the fuel economy
advantage of a higher-powered diesel vehicle would be expected to be smaller
than that for current lower-powered diesel vehicles.
S-31
-------
Table S-8. Average Light-Duty Diesel Vehicle Emissions
(1975 Federal Test Procedure)
Test Vehicle
Mercedes Benz Z20D
Mercedes Benz 220D
(modified injection)
Mercedes Benz 220D
Opel Rekord 2100D
Peugeot 504 Diesel
Peugeot 504 Diesel
Ford/Nissan Diesel
Test
Laboratory
EPA
EPA
Daimler-Benz
EPA
EPA
Peugeot
EPA
Emissions, gm/mi
HC
0.34
0.28
0. 15
0.40
3. 11
2. 50
1.70
CO
1.42
1.08
2.3
1.16
3.42
2.20
3.81
NO
X
1.43
1.48
1.45
1.34
1.07
1.30
1.71
gm/mi = grams per mile
S-32
-------
Table S-9. Light-Duty Diesel and Gasoline Vehicle Fuel Economy
Comparisons (1975 Federal Test Procedure)
Diesel Powered
Vehicle
Mercedes Benz 220D
{standard injection)
Mercedes Benz 220D
(modified injection)
Opel Rekord 2100D
Peugeot 504
Ford/Nissan Pickup
Fuel
Economy,
mi/ gal
23.6
24.6
23.8
25.2
21.4
Gasoline Powered
Vehicle
1973 Mercedes 220
1973 Mercedes 220
1973 Peugeot 504
1973 Average
3000-pound car
. 1973 Average
3500-pound car
1973 Average
4500-pound car
Fuel
Economy,
mi/ gal
13.2
13.2
17.0
15.6
13.9
10. 1
mi/ gal = miles per gallon
S-33
-------
Power output per pound of engine weight is considerably
lower than that presently offered in spark ignition engine-powered domestic
automobiles. Methods for increasing the power output per pound of engine
weight include (a) means to increase power by turbocharging or supercharging
and (b) means to reduce engine weight and bulk by a reduction in engine com-
pression ratio and a shortened design life.
S.I.5 Wankel and Other Rotary Piston Engines
S.I.5.1 General Description
In its simplest form, the Wankel engine consists of a single
triangular-shaped rotor which revolves eccentrically in a chamber with a
double-lobed cross section (Figure S-14). The motion of the rotor is such
that the apexes of the rotor continuously brush the walls of the chamber,
thereby forming three compartments which rotate around the chamber,
changing size and shape as they move by virtue of the eccentricity of rotation.
Thermodynamically, the engine is based on the Otto cycle in its "four-stroke"
form. On one side of the chamber casing is a spark plug. Each compart-
ment passes successively through all phases of the Otto cycle once in each
rotor revolution. Since there are three compartments, three power strokes
are produced with each rotor's revolution.
S. 1.5.2 Historical Development
Many efforts have been expended in the past century to
develop internal combustion engines that will produce a torque at a reasonably
low speed without need for a crankshaft and its attendant complications.
Within the last decade or so, these efforts have produced a wide variety of
prototype rotary piston engines, largely experimental in nature.
Numerous rotary engine design concepts have been proposed,
but the one receiving the greatest attention for automotive application is
the Wankel Rotary Piston Engine invented by Felix Wankel in 1953. The
prototype was run successfully at the NSU Works in Germany on 1 February
1957. Since then, experimental engines of various sizes have been built and
S-34
-------
AIR INTAKE
EXHAUST
a.
SPARK
PLUG
COMPRESSION
b.
c.
POWER
EXHAUST
Figure S-14. Wankel Rotary Engine
S-35
-------
tested for a wide variety of applications covering automobiles, aircraft,
boats, snowmobiles, lawn mowers, and auxiliary power units. Current
automotive rotary piston engines are water-cooled and are designed with one
or two rotors for production cars and up to four rotors for experimental
models. A list of vehicles equipped with these engines, together with perti-
nent data, is given in Table S-10.
S.I.5.3 Current and Project Status
Although currently a very small fraction of the automobile
population, rotary engine-powered cars have been rising in production
levels. The Chevrolet Vega with the General Motors rotary engine is
scheduled for limited introduction in the 1975 model year.
Rotary engine-powered car production levels have been rising,
although they represent a very small fraction of the automobile population.
Currently, these engines appear principally in Mazda vehicles manufactured
by Toyo Kogyo. Production levels are expected to increase with the limited
introduction scheduled for the 1975 model year of the Chevrolet Vega
powered by the General Motors rotary engine.
A manufacturing problem exists in attempts to increase pro-
duction rates. Because of complex machining operations required on both
the housing and rotor, production of this engine is currently limited to about
25 per hour per line (compared to piston engine rates of over 100 per hour
per line). This limitation appears to have a significant impact on manufac-
turing costs.
Technological problems still exist for some engine designs.
These consist of high gas seal leakage, high thermal stresses, poor low-
speed fuel economy, and high exhaust emissions without aftertreatment.
With aftertreatment (a thermal reactor), the Mazda rotary engine has demon-
strated the ability to meet the original 1975 Federal emission standards
(Table S-ll).
S-36
-------
Table S-10. Vehicles Equipped with Rotary Piston Engines
Vehicle
NSU Wankel
Spider
NSU Ro-80
Mazda Cosmo
110-S
Mazda R-100
Mazda RX-2
Citroen M-35
Mercedes Benz
C-IH MKI
Mercedes Benz
c-m MK n
Mustang RC 2-60
Introduction
1964
1967
1967
1969
1970
1970
1969
1970
1965
Discontinued
1967
—
1969
—
—
Limit 500
cars
Expe rimental
Experimental
Experimental
Engine
NSU
NSU
Toyo Kogyo
Toyo Kogyo
Toyo Kogyo
NSU
Mercedes Benz
Mercedes Benz
Curtiss- Wright
Horse-
power
64
136
110
110
120
55
330
400
185
Rotors
1
2
2
2
2
1
3
4
2
Compression
Ratio
8.6:1
9.0:1
9.4:1
9.4:1
9.4:1
9. 0:1
9.3:1
9.3:1
8.5:1
cn
OJ
-v]
-------
Table S-ll. Best Emissions Results, Thermal Reactor and Wankel
Engine, 2750-Pound Compact Car
C/)
i
OJ
00
Manufacturer
Toyo Kogyo
Toyo Kogyo
General
Motors
System
Reactor
Reactor
Reactor +
EGR
Reactor
(Best effort)
Mileage
4,000
50, 000
Low
Low
Low
No. of Tests
1973
Certification
Many
3
8
72
1
Emissions ,
gm/mi
HC
2.4
0.32
0.36
0.35
0.60
0.43
CO
20.0
3. 1
2.6
2.2
5.0
2.8
NO..
.A.
0.9
0.83
0.87
0.49
0.60
0.44
FuelC
Penalty,
%
0
6.5
--
12.0
--
All these systems are carbureted fuel rich (air -fuel ratio of about 12).
1975 Constant volume sampling, cold/hot test procedure (grams per mile).
Relative to 1973 production rotary engine car.
-------
S.I.6 Stratified Charge Engines
S.I.6.1 General Description
In principle, the stratified charge engine represents a modi-
fication of the conventional spark ignition engine. The principal objective of
the modification is the achievement of heterogeneous combustion in which a
rich fuel-air mixture is generated around the spark plug and a lean mixture
in the remaining zones of the combustion chamber. The resulting two-stage
combustion process permits operation of the engine at very lean overall air-
fuel ratios that are conducive to low emissions, good fuel economy, and
reduced sensitivity of the engine to fuel octane and cetane numbers.
The stratified charge engines can be divided into two distinct
classes: open-chamber engines (Figure S-15) and divided-chamber engines
(Figure S-16). In the open-chamber configurations, exemplified by the
Texaco TCCS and Ford PROCO engines, a single combustion chamber is
used similar to that of conventional spark ignition engines. Upon ignition of
a swirling, rich mixture surrounding the spark plug, the burning charge
expands into the power regions of the combustion chamber where the com-
bustion process is then completed in an oxygen-rich environment.
The divided-chamber stratified charge, or prechamber,
engines, exemplified by Honda's CVCC engine concept, employ two inter-
connected combustion chambers per cylinder. During the compression stroke
of the piston, a fuel-rich mixture is inducted into the generally smaller pre-
chamber while the main chamber is charged with a lean mixture or even with
pure air. Upon ignition in the prechamber, hot gases expand into the main
chamber where combustion is then carried to completion. The principal
advantage of prechamber engines over conventional engines is their ability
to operate with very lean overall fuel-air mixtures resulting in concurrently
low hydrocarbon, carbon monoxide, and oxides of nitrogen emissions.
S-39
-------
CO
•^
o
Figure S-15. Texaco Cup Combustion,
Open- Chamber
Stratified Charge Engine
Figure S-16. Honda CVCC,
Divided-Chamber
Stratified Charge Engine
-------
S.I.6.2 Historical Development
The concept of stratified charge operation of internal
combustion engines and its potential was first evaluated in the early 1920's
by Ricardo in England. In the United States, work on open-chamber con-
figurations commenced in the 1940's with Texaco's invention of a "knockless"
engine. In 1965, the United States Army Tank Automotive Command
(TACOM) initiated funding to Texaco and Ford for exploratory research of
both open-chamber and divided-chamber stratified charge engine concepts
directed toward developing a military engine with multifuel capability and
improved fuel economy relative to conventional spark ignition engines.
In 1970, the relatively low exhaust emissions of the available
stratified charge engines and the apparent potential for improvement led to
additional funding by EPA through the Army. The TACOM program was a
joint Army/EPA-funded effort. Figure S-17 shows the AAPS Program sched-
ule for stratified charge engine development. Analytical and test evaluations
of the manufacturing and operational (performance, reliability, maintain-
ability) aspects of the current designs were emphasized to aid in preparing
improved design specifications for the next generation of engines.
The concept of a divided combustion chamber can be traced
back to the first "oil engines" in the pre-Diesel era. These early pre-
chamber engines operated on heavy oils or naphtha, which was injected into
a spherical prechamber and vaporized in contact with the red-hot thermally
insulated prechamber walls. Later, attempts were made to apply the pre-
chamber concept to the spark ignition gasoline engine, as documented by
numerous patents. However, to date a practical widespread application
of the prechamber concept has been realized only in high-speed diesel
engines, principally due to the pioneering efforts of Ricardo. A number of
organizations and individuals are currently working to develop prechamber
S-41
-------
PROGRAM PHASE
CALENDAR YEAR
1972
1973
1974
HARDWARE EVALUATION
(4-cylinder engineering
prototypes)
ENGINEERING EVALUATION
ENGINE DEVELOPMENT
(production prototypes)
BUILD AND INSTALL IN
PASSENGER CARS
J
EPA TEST
MANUFACTURING/INTEGRATION/OPERATION
EVALUATION
L
J
STATISTICAL ANALYSES
J
DESIGN IMPROVEMENT
I
L
J
BUILD SOFT-TOOLED ENGINES
I I
L
J
EMISSIONS, PERFORMANCE AND DURABILITY TESTS
Figure S-17. Stratified Charge System Army /EPA Development Program
-------
engines. In particular, during the past several years, the Honda Motor
Company of Japan has conducted extensive design, development, and test
work on its CVCG concept.
S.I.6.3 Current and Projected Status
Stratified charge engines developed to date, notably by Ford,
Texaco, and Honda, have achieved an advanced state of development.
Without incorporation of additional emission control systems,
the hydrocarbon and nitrogen oxide emissions from TACOM M-151 vehicles
adjusted for best fuel economy and equipped with experimental Ford and
Texaco open-chamber stratified charge engines are comparable to those
from vehicles powered by conventional spark ignition engines, while carbon
monoxide emissions and fuel economy are substantially improved. Con-
versely, -with emission control (consisting of exhaust gas recirculation, oxi-
dation catalysts, and intake air throttling at low power levels), the vehicles
meet the original 1976 Federal emission standards at low mileage and show
equal or slightly better fuel economy than the average 1973 spark ignition
engine -powered certification vehicles tested at equivalent inertia weights.
But these M-151 vehicles were unable to negotiate the high acceleration
modes of the Federal Emission Test Driving Cycle. Table S-12 presents
comparative data for the M-151 vehicle equipped with the Ford PROCO strati-
fied charge engine; similar comparative data are shown in Table S-13 for the
Texaco TCCS engine.
Honda has a divided-chamber version (designated CVCC) of
the stratified charge engine in production. Without incorporation of addi-
tional emission control systems, its CVCC engine-powered Civic vehicle
met the original 1976 Federal emission standards for HC and CO, while
NO was about twice the standard (Table S-14). Vega and Impala vehicles
achieved similar emission levels at low mileage when powered by General
Motors engines modified by Honda to the CVCC configuration.
Fuel economy of the Honda Civic vehicle with the CVCC
engine, as measured over the Federal Emissions Test Driving Cycle, was
S-43
-------
Table S-12. Emissions and Fuel Economy, M-151
PROCO Vehicles
Vehicle
PROCO, no catalyst
PROCO, with catalyst
PROCO, with catalyst
Emissions,
gm/mi
HC
2. 60
0. 35
0. 37
CO
13.45
1. 01
0.93
NOX
0. 32
0.35
0. 33
Fuel
Economy, U)(4)
mi /gal
21. 7
21.3(2)
Not measured
Number of
Tests
(averaged)
1
4
14
Test
Facility
Ford
Ford
EPA
197Z CVS Test Procedure
PROCO, best fuel
economy, no catalyst
PROCO, no catalyst
PROCO, with catalyst
4. 96
3. 10
0. 54
7. 75
13. 75
1. 18
3. 85
0.33
0. 37
23. 8
21.2
19.6
2
1
4
Ford
Ford
Ford
CVS = constant value sampling.
(1) Computed from the mass emission data (miles per gallon).
(2) Z0.4 miles per gallon, measured with a burette.
(3) Baseline (carbureted) vehicle emissions are 4. 55 gm/mi HC, 41.6
gm/mi CO, 4.4 gm/mi NOX.
(4) Baseline (carbureted) vehicle fuel economy 17.2 mi/gal.
about ten percent lower than that of conventional Civic vehicles, and
16 percent lower than that of equivalent-weight 1973 model year emission
certification vehicles. A Vega vehicle with this type of engine showed fuel
economy five to ten percent better than that of the standard Vega. A CVCC
powered Impala and a Texaco TCCS-powered M-151 vehicle showed equal or
slightly better fuel economy compared to the respective unmodified vehicles.
Both open-chamber and divided-chamber stratified charge
engines have demonstrated a lower sensitivity to fuel octane number than
the basic spark ignition engine. Because of the late fuel injection combined
with immediate spark ignition, the Texaco TCCS engine can operate also on
a wide range of diesel fuels.
S-44
-------
Table S-13. Effect of Emission Control System Modifications on M-151 TCCS Vehicle
Emissions and Fuel Economy (1975 Federal Test Procedure)
en
Engine Configuration
Tested at Texaco
Max. economy setting, no catalyst
Max. economy setting, no catalyst
8 deg. combustion retard only, no catalyst
8 deg. combustion retard, low EGR rate
8 deg. combustion retard, medium EGR
rate, two platinum catalysts
13 deg. combustion retard, high EGR rate
two platinum catalysts
5 deg. combustion retard, no EGR, two
platinum catalysts, alternate vehicle
Baseline configuration
Baseline configuration, with reduced EGR
Baseline configuration '
b c
Baseline configuration, ' with reduced EGR
Turbo-
charged
Yes
No
Yes
Yes
Yes
Yes
Yes
No
No
No
No
Emissions, a
gm/mi
HC
3. 13
4. 24
3. 24
3. 60
0. 33
0. 35
0. 30
0. 36
0. 48
0. 37
0. 50
CO
7. 00
7. 28
6.43
6.69
1. 05
1. 41
1. 07
0. 61
0. 57
0. 24
0. 14
NOX
1.46
1.43
1. 29
0.84
0.61
0. 35
1. 40
0. 31
0. 45
0. 31
0. 70
Fuel
Economy,
mi/gal
24. 3
--
22. 4
20. 5
19.7
16.2
20. 9
16. 2
17 6
15. 8
21. 9
Fuel: gasoline, 91 research octane EGR = exhaust gas recirculation
number, plus 2% oil. mi/gal = miles per gallon (by weight)
b_,,~ , , .. . , gm/mi = grams per mile
EGR plus two platinum catalysts. " 6 c
CTested at EPA.
-------
Table S-14. Emissions and Fuel Economy, Honda
Civic CVCC Vehicles
Vehicle
Low-mileage car
(#3652. Five tests.)
Average
Maximum
Minimum
Low-mileage car
(tf3606. One test.)
50, 000-Mile car
(fl<2034. Four tests.)
Average
Maximum
Minimum
Original 1975 Federal
standards
Original 1976 Federal
standards
Emissions ,
gm/mi
HC
0. 18
0.21
0. 15
0.23
0.24
0.26
0. 19
0.41
0.41
CO
2. 12
2.28
1.96
2.00
1 .75
1 .85
1 .70
3.40
3.40
NOX
0.89
1.05
0.75
1 .03
0. 65
0.73
0. 57
3. 1
0.40
Fuel Economy, mi/gal
1975 FTP
22. 1
22.4
21.9
20.7
21 .3
22.2
20.8
1972 FTP
21.0
21. 5
20.6
19.5
19.8
20.0
19.5
a!975 Federal Test Procedure (FTP) (grams per mile).
The values of the different columns do not necessarily correspond
to the same test.
mi /gal = miles per gallon
S-46
-------
The production cost of open-chamber stratified charge
engines is expected to be higher than that for conventional engines because
of the required fuel injection system, and for the divided-chamber version,
a small cost increase over conventional engines is indicated because of the
more complex cylinder head configurations.
The open-chamber stratified charge engines fabricated to date
are experimental models not suitable for mass production. A number of poten-
tial problem areas required further evaluation, including emission control sys-
tem durability, and the current inability to achieve oxides of nitrogen emission
levels that meet original 1976 Federal emission standards without incurring a
substantial loss in fuel economy. Ford feels that limited production of PROCO
engines might be feasible by 1977. Conversely, the prechamber configuration
developed by Honda is currently in limited mass production.
S.2 COMPARATIVE REVIEW OF ALTERNATIVE
HEAT ENGINES
Each of the engines being considered herein as an alternative
to the conventional spark ignition engine for automobile propulsion has its own
unique characteristics. If adequate performance data were available, the
engines could be compared to provide a relative ranking of these character-
istics. In general, such data are limited in scope or not available for this
purpose, because most of these engines are in an early state of development
where more effort has been devoted to establishing proof-of-principle for a
particular design concept than to establishing all of the prototype baseline
performance characteristics. Therefore, in most cases, comparisons are
at best subjective.
Of overriding importance, however- is that each alternative
engine offers the potential of meeting the original 1976 Federal emission
standards, and most show reasonable progress toward these standards as
shown by the representative data for selected engines in Table S-15. It
should be stressed that because the engines are at different stages of tech-
nological development, it is not meaningful to attempt to compare and rank
them with regard to the potential for achieving the lowest possible emissions,
S-47
-------
Table S-15. Selected Exhaust Emission Data, Alternative Heat Engines
C/)
i
4x
oo
Engine
Gas Turbine
Chrysler engine
Solar advanced
combustor
General Motors
2 25 -horsepower
Rankine Cycle
Water Reciprocating
Stirling
Diesel
Mercedes Benz 220D
Wankel
Toyo Kogyo
Stratified Charge
Honda CVCC
Ford M-151 PROCO
Exhaust Emissions ,a
gm/mi
HC
-0. 26
0. 11
0.02
0. 09
0. 10
0.34
0.35
0.24
0.37
CO
3.99
0.9
2.4
0.5
0.31
1.42
2.2
1.75
0.93
NOX
2. 21
0. 17
0.32
0.2
0. 18
1.43
0.49
0.65
0.33
Remarks
Background level subtracted
from measured values
Calculated, based on
combustor tests
Chassis dynamometer
tests, experimental engine
Calculated, based on
component tests
(1972 FTP)
Calculated, based on
component tests
EGR and engine modifi-
cations required to
reduce NO emissions
Low-mileage tests with
thermal reactor and EGR
Average values for 50, 000-
mile test car (Civic)
Based on EPA tests with
catalyst and EGR
a!975 Federal Test Procedure (FTP) EGR = Exhaust gas recirculation.
except where noted (grams per mile).
-------
whether for a single exhaust emission specie (e.g. , hydrocarbons) or all
three species.
Besides emissions, other topics of concern in selecting an
engine to supplant the spark ignition engine are: adaptability to mass pro-
duction manufacture, purchase and maintenance costs (including durability
factors), fuel consumption, fuel compatibility, weight and size impact on
vehicle packaging requirements, flexibility and responsiveness to driver
commands, noise, etc. Current development efforts are directed toward
engine size and weight reductions that will permit engine installation in
conventional automobile engine compartments. Additional effort is also
applied to ensuring engine flexibility and responsiveness and to reducing
noise levels.
Mass production manufacturing and the resulting purchase
cost are factors that have been reviewed in screening the various alternative
engines, but no firm quantitative comparisons are available except in a few
specific cases. According to a production implementation study conducted
for the U.S. Department of Transportation, dual-shaft, regenerative gas
turbines are estimated to cost about twice as much as current spark ignition
engines (estimated retail costs ranged from about $550 to $650 for a spark
ignition engine). Under low production rates, the cost of current diesel-
powered cars ranges from approximately $200 to $700 more than current
spark ignition engine-powered cars as cited in congressional hearings con-
ducted in 1973 with respect to suspension of the 1975 Federal emission
standards. To provide a proper basis of reference, however, there must be
an accounting of the added cost to the consumer of exhaust emission control
equipment for the spark ignition engine designed to meet 1976 Federal
emission standards. Based on a study conducted for EPA on the effects of
lead additives in gasoline on emission control systems for 1975-1976 motor
vehicles, these costs are expected to range from approximately $200 to $400.
With respect to fuel consumption, the data are sufficient to
afford an illustrative comparison among various engines. Measured fuel
S-49
-------
economy results for spark ignition engines, contrasted with measurements
and estimates of fuel economy for alternative engines, are presented in
Figure S-18. As can be seen, some of these alternative engines appear
capable of delivering superior fuel economy, particularly when judged
against the degraded fuel economy of exhaust-emission-controlled 1973
model year cars. Even on this basis, an absolute comparison is difficult
because the fuel economies of the internal combustion engines change with
model year.
The compatibility of each engine with various types of fuels
can be a strong factor in its selection, particularly for the period beyond
1985 when advanced engines combined with synthetic fuels may be considered
practical and significant contributors to national personal transportation.
25
|
8>20
k.
8.
I15
>-
o
o
u
10
SU 5
0 MEASURED ENGINE IN VEHICLE (NOX - 0.4 gm/mi)
O MEASURED ENGINE IN VEHICLE (NOX - 2.0 gm/mi)
A PROJECTED BASED ON ENGINE COMPONENT MEASURED
PERFORMANCE (NOX - 0.4 gm/ml)
D PROJECTED DATA MODIFIED FOR POWER TO WEIGHT RATIO
OF SPARK IGNITION ENGINE POWERED CARS
gm/ml = grams per mile
O
HONDA
CVCC
TEXACO
STRATIFIED
CHARGE
(jeep)
WANKEL CARTER
MAZDA STEAM
RANKINE
WILLIAMS O
RESEARCH
GAS TURBINE
SPARK IGNITION ENGINES
AVG 1957-1967 CARS
(NOx-6.0 gm/mi)
I I I I
STIRLING
GAS
TURBINE
(NOX~3.1 gm/mi f
I I
2000 25000 3000 3500 4000 4500
INERTIA WEIGHT, pounds
5000 5500
Figure S-18. Fuel Economy Over the Federal
Emissions Test Driving Cycle
S-50
-------
Those engines utilizing continuous combustion processes (e.g. , gas turbine,
Rankine cycle, Stirling) offer the widest latitude in the use of alternative
fuels. The intermittent combustion process engines are somewhat more lim-
ited in flexibility, particularly the diesel engine which requires fuels of greatly
restricted range in cetane number to sustain combustion in the compression
ignition process.
An estimate of the current state of development and the overall
required development time for each alternative engine is important in assess-
ing the period when such an engine might be available in large quantities as a
marketable automobile power plant. On the basis of information assembled in
a study of gas turbine engine production implementation prepared for the U.S.
Department of Transportation, Figure S-19 shows the estimated product devel-
opment time for a new automobile engine. By assuming similar development
time for various types of engines, one estimate of the current point of progress
for each of the alternative engines under discussion has been recorded on the
time scale. It should be noted that the pace of progress from this time onward
is not necessarily the same for each engine. It is dependent on the assigned
development funding levels and on the complexity of problems encountered in
efforts made to meet engine performance and manufacturing design require-
ments. The Stirling engine lags behind all other engines and, were it to be
*
mass produced, it is estimated to be almost nine years from full production.
0< O*
The rotary engine is only in limited production at this time
by Mazda. (General Motors plans to introduce this engine in limited quanti-
ties for the 1975 model year Vega.)
Diesels are shown as being in limited production. Mercedes
Benz has a low performance diesel in production, and Peugeot has started
production in model year 1975. However, a lightweight, high-performance
engine for automobiles is still in a research prototype development phase and
is estimated to be about 7 years away from full production.
Approximately 300, 000+ units per year
**
Approximately 20, 000 units per year
S-51
-------
PROGRAM PHASE
CONCEPT DEVELOPMENT
(Engine-Transmission-Vehicle)
RESEARCH PROTOTYPE
DEVELOPMENT
ENGINE PILOT PLANT
INSTALLATION
PROVEOUT AND STARTUP
VEHICLE PROTOTYPE OR
LIMITED PRODUCTION
FLEET TESTS
ENGINE MASS PRODUCTION PLANT
INSTALLATION
PRODUCTION BUILDUP
VEHICLE FULL PRODUCTION
YEARS FROM PROGRAM START
1
2
3
4
5
I
^^^ STIRLING-
|HB PRODUCTION DESIGN RAN
[Illllllllllllllll PRODUCTION DEVELOPMENT ^Til
VJIF"
(ope
J
KINE '
ATIFIED —
RGE
n chamber)
6
1
•
7
8
1
1
Illlllliil
1
GAS
TURBINE
-DIESEL
(high pert)
-STR/
(prec
9
10
I
-DIES
ROT
kTIFIEl
:hambe
*
•
11
12
•Hi
HUH
_
EL (lowperf)
ARY (Wankel)
3 CHARGE
r)
Figure S-19. Automobile Mass Production Implementation Schedule
for a New Engine
-------
The prechamber version of the stratified charge engine is
considered to be in a limited production phase on the basis of Honda's
introduction of their CVCC engine in the 1975 model year Civic. But Ford
feels that their open-chamber design could only be produced in limited
numbers by 1977.
If gas turbines were to be mass produced, it is estimated
that these engines are about 6-1/2 years away from full production, followed
by Rankine cycle engines at almost 8 years from full production. The
proximity in time of the gas turbine engine to pilot plant installation is based
on the limited production of engines that Chrysler carried out during their
50-car program, the experience gained by General Motors in small produc-
tion of turbines for trucks, and the experience of Solar and Ford in carrying
out the production design effort for limited production of turbines for trucks.
At this time no alternative engine development appears to have
progressed to a point where sufficient data are available to substantiate any
claim of superiority over other engines. Although the engines discussed in
this report had been formerly viewed with the prospect of meeting the
original 1976 Federal emission standards, the recent emphasis on con-
servation of natural resources has prompted a review of engine performance
to promote high fuel economy and wide fuel adaptability in concert with
low exhaust emissions. This trend may be seen in the renewed interest
in the Stirling engine, whereby the Philips and United Stirling lightweight
designs are being actively investigated by Ford for prototype installation
in Torino and Pinto cars, respectively.
S-53
-------
SECTION 1
-------
1. INTRODUCTION TO HEAT ENGINE SECTION
Enactment of the Clean Air Act Amendments of 1970 triggered
vigorous research efforts on the part of the automotive and related industries
(* **)
to find the means to meet the original ' 1976 exhaust-emission standards
for hydrocarbon (HC), carbon monoxide (CO), and nitrogen oxides (NO )
2C
(0.41, 3.4, and 0.40 gm/mi, respectively) required for the 1976 model year
automobile and beyond. Most of this industry effort was focused on attempts
to clean up the conventional spark ignition internal combustion engine (ICE).
Some of the research money went into investigating the feasibility of low-
emission types of engines as an alternative to the conventional ICE.
The ability of the conventional ICE to meet the 1976-1977
Federal exhaust emission standards has been in doubt for some years. Even
with the most elaborate emission control schemes, the prospects for meet-
ing these standards are considered marginal by most auto makers. In
addition, the add-on emission control systems are expensive, they affect
drivability of the car, their durability for 50,000 miles is a serious problem,
and they result in degraded fuel economy.
As an alternative, a number of other engines have been con-
sidered for the automobile power plant of the future. These engines, as
conceived, may not require add-on devices. In most cases, it is expected
that the basic combustion process can be adequately controlled to achieve
the expected low emission levels. In this regard, the most promising
^ Interim standards with relaxed requirements are now in effect.
{C
Exhaust emission goals for engines under investigation in the AAPS Program
are set at one-half the Federal standard. Requirements are made more
severe to: (a) account for changes from prototype to mass-production designs,
(b) ensure that production engines, with expected statistical variations, can
pass the Federal certification test, and (c) allow for expected increases in
emissions as vehicle mileage reaches the 50, 000-mile Federal certification
limit.
1-1
-------
systems rely on continuous combustion processes as opposed to the intermit-
ent combustion processes found in a spark ignition engine. However, low ex-
haust emission alone cannot be the sole basis for selecting an engine of the
future. These advanced engines, when ranked against equivalent horsepower
spark ignition engines, must also fulfill a number of other requirements,
among which are:
a. Competitive fuel economy both in stop-and-go traffic
as well as at highway cruise speeds
b. Physical dimensions compatible with automobile engine
compartments
c. Ability to operate satisfactorily under a wide range of
climatic and altitude conditions
d. Competitive first cost and maintenance cost
e. Adaptability to mass production manufacturing processes
f. Competitive durability
g. Adequate responsiveness to driver commands.
Evaluation of low-emission alternative engines to meet these
requirements has been a continuing task since 1970 when a program of
research and technology development was launched by EPA to evaluate the
feasibility of various automotive power plants. The program is called the
Alternative Automotive Power Systems (AAPS) Program and is administered
by the Alternative Automotive Power Systems Division of EPA. This
program was intended to supplement the work of industry, thereby ensuring
that all reasonable technical approaches to meeting the standards were being
pursued. A few of the engines under study were invented several decades ago
and have been used successfully as power generators for a number of years.
However, until the recent application of modern-day technology, these con-
cepts were considered too inadequate in weight, size, cost, and efficiency for
serious consideration as a mass-produced automobile power plant.
In July 1971 it was decided to limit the number of candidate
systems under development and to concentrate all available funding and man-
power resources on the three most promising systems for the near term.
1-2
-------
These systems are the gas turbine, Rankine cycle, and stratified charge
engines. The theoretical characteristics and anticipated development
schedules of these three systems were considered more favorable than the
others. Consequently, development work is focusing on these three, with
major efforts going to the gas turbine and Rankine cycle engines because of
their greater need for new technological advancements.
By means of hardware demonstration of alternative power
systems, it is hoped that the AAPS Program will stimulate industry to absorb
technologies developed in the program and further develop the candidate sys-
tems into production engines. Since its inception it has been apparent that
the program has served additionally as an important mechanism for trans-
ferring aerospace technologies developed under federal government funding
in NASA and DOD programs.
This volume of the report to the Administrator is intended to
provide a summary of the status of engine development work in this program
as well as in similar alternative engine hardware development programs
being funded by private industry. Operating characteristics of the alterna-
tive systems are also provided. Still more information on other developments
is known to the government but cannot be documented herein because the data
are proprietary. Finally, the status of developments as of the end of 1973 is
summarized, wherever possible, in the report.
One purpose of the AAPS Program serves to provide a basis
for knowing what can and cannot be accomplished with alternative engines
used in practical ways; thus, the publication of such information as contained
herein is in keeping with the mission of the AAPS Program mission. It is
intended that this series of reports on alternative engines be published annu-
ally, with the present report being the first. The program is scheduled to be
completed in 1975 with demonstrations, testing, and evaluations of Rankine
cycle and gas turbine engines installed in automobiles.
1-3
-------
SECTION 2
-------
2. OPERATIONAL REQUIREMENTS AND CHARACTERISTICS
OF AUTOMOBILES POWERED BY CONVENTIONAL
SPARK IGNITION ENGINES
This section of the report briefly addresses and summarizes
the salient features and characteristics of automobiles powered with con-
ventional spark ignition engines. It is included herein in order to provide
a basis for perspective and comparison between the conventional spark
ignition engine and the various alternative automotive engines examined in
detail in later sections of this volume—in particular, with regard to vehicle
performance, fuel economy, and exhaust emissions.
Initially, a brief historical review of the evolution of the
spark ignition engine-powered automobile is presented in Section 2. 1. Next,
automobile systems and operational requirements are discussed with respect
to the interrelationships influencing vehicle performance capability and fuel
economy (Section 2.2). Finally, Section 2.3 summarizes exhaust emission
control systems applicable to the conventional spark ignition engine and
resultant exhaust emissions as characterized by current federal exhaust
emission standards. Appendix A contains a listing of contemporary vehicle
design specifications as provided by the AAPS Program to form a reference
vehicle for comparisons of alternative engine performance when installed
in the automobile.
2. 1 EVOLUTION OF THE SPARK IGNITION ENGINE*
2.1.1 Early History of Auto Engines
The task of determining why the internal combustion is the
engine of primary use is as difficult as determining who discovered America.
*
Selected abstracts from Reference 2-1
2-1
-------
The historical record of locomotive power is an exercise in love and
confusion. The affection of authors and historians for one or another of
the electric, steam, or internal combustion engines results in the favored
machine being bathed in rose-colored lights. Confusion exists because few
sources -- including the papers by the principals themselves -- agree even
on dates, let alone on the factors that influenced the ultimate success or
failure of the power unit.
Nevertheless there are numerous identifiable factors, indi-
viduals, and circumstances involved in the trail that leads ultimately to
our current automobile. Among these are the political, sociological,
economic, technological (internal and external), and institutional forces
that influenced the decisions of individuals and firms. Personal drives of
engineers, promoters, and financiers were also instrumental in shaping
the current automobile industry.
The critical period for the fledgling auto industry was from
about 1900 to about 1920:* The period *** was *** one of ferment and
growth for the entire automobile industry. That activity was in fact a phan-
tasmagoria of bold ideas, ingenious and stubborn experimentation, enthu-
siastic promotion, dramatic accomplishment, and weird display. From it
developed the steadier era of manufacturing in which Henry Ford would
emerge as a leader.
Although by 1900 at least 57 American plants were engaged in
making motor cars, many of which were in an experimental state, the basic
type of vehicle had not yet been determined, and it was uncertain where the
chief centers for the new industry would develop. This year saw about 4,000
American cars manufactured, of which steam and electric models made up
more than three-fourths. One respectable set of figures lists 1,681 steam
carriages, 1,575 electric, and only 936 gasoline cars.
*Abstracted from Reference 2-2.
2-2
-------
2.1.2 Factors Affecting Early Decisions in Favor of
Spark Ignition Engines
The literature of the late 19th and early 20th centuries, as
well as more recent studies of that literature and of that period, does not
provide the answer to the question as to why the spark ignition engine
became the engine of choice. However- that literature does support the
following conclusions:
a. At the time of its original selection as the engine of
choice, the spark ignition engine had no overall intrinsic
advantage over other types of engines.
b. The spark ignition engine became the engine of choice
as the result of a number of circumstances and events
rather than as the result of a formal decision-making
process. Among the circumstances and events were:
1. Although steam-powered road vehicles were
developed in the 1820s, restrictive legislation in
England in force between the 1830s and 1896
discouraged further developmental efforts during
the period of conception and early development
of the spark ignition engine.
2. Henry Ford picked the spark ignition engine to
power his cars. When the Model T succeeded, the
spark ignition engine was assured of at least a
place in the range of auto engines to be used. Why
Ford picked the spark ignition engine is not clear,
beyond his great concern for light weight.
3. Once Ford was in mass production, his competitors
had to pick a basis on which to compete against him.
His most successful competitor, General Motors,
decided to compete on the basis of financial and
marketing values and not on the basis of engine type.
4. Standard Oil Company, with its development of the
Burton process to crack crude oil, assured a plenti-
ful supply of inexpensive, though "highly refined,"
gasoline needed by the spark ignition engine but by
none of the other types.
5. Adding the electric starter, which removed the
problem (and danger) of hand-cranking to start the
engine, opened the women's market to spark ignition-
powered cars. Up to that point, women had preferred
the electric car.
2-3
-------
6. When Doble entered into competition against Ford
with a steam-powered car, World War I came
along, causing the Government to intervene in the
interest of war production. Doble did not fill
$27 million in standing orders; when the war was
over any momentum that might have built up for
Doble's steam-powered competitor was gone
forever.
7. In the early 1920s, when roads were becoming
good enough to support a demand for higher engine
performance, Midgley discovered tetraethyl-lead
which, when added to gasoline, provided the spark
ignition engine with the performance capability to
defend its position against the known high-speed
performance of the steam-powered engine.
c. As car buyers increased the stringency of their demands
for engine performance, the industry successfully modi-
fied the spark ignition engine to satisfy those demands.
There has been no Achilles heel in its performance on
which competitive types of engines could base a
challenge.
2.1.3 Factors Influencing Continued Reliance on Spark
Ignition Engines
Fierce arguments raged for a quarter century or more over
whether the spark ignition engine, the steam engine, or the electric battery
would be the best way to power a car. Reading those arguments today, one
cannot help but wonder at the willingness of any entrepreneur to risk his all
on any one of them. But at least he didn't have to worry about displacing
an engine already performing satisfactorily, as becomes apparent when the
following factors are considered:
a. In contrast to the situation in 1900, the spark ignition
engine today has a fundamental intrinsic advantage
over all other candidates. Over more than 50 years,
hundreds of millions of them have been made, used,
maintained, and repaired:
1. Massive production facilities, tools, and methods
exist for their large-scale production.
2-4
-------
2. Designers have an immensely detailed and
documented understanding of factors affecting
their design, performance, and cost.
3. Hundreds of thousands of people are trained in
their maintenance and repair.
4. A network of service facilities, with tools and
spare parts distribution, has been built up.
5. Millions of customers through personal
experience have gained confidence in spark
ignition engines.
6. There is incontrovertible evidence that progress
is currently being made in designing them to
operate at lower emission levels without intolerable
sacrifices in other parameters of performance.
There are no equivalents in knowledge, experience, or
data with alternative engine designs. Thus it is not
possible to prove that a change to one of these alterna-
tive designs would be beneficial --in either emission
reduction or other performance parameters.
Reducing emissions is not the only change in engine
design parameters that must be incorporated over the
next decade. Fuel source, fuel economy, and transpor-
tation functions are in process of reconsideration as a
result of major shifts in population, social values, and
resources. Many of the factors involved in these
changing parameters are not yet clear. In light of such
uncertainty, decision makers seek to keep open as many
options as possible for as long a time as possible or,
conversely, to make short-term decisions as incremental
as possible and to postpone major "change" decisions.
All technological innovation, including incremental steps,
follows a pattern that can be expected to apply to the auto
industry. Typically, an innovation is introduced on a
small scale, tested, and proved; gradually it penetrates
the market. The period of experimentation varies widely,
depending on numerous factors, including those listed
above in a and b_, and involving the complexity of the
innovation, the extent of the supporting system for the
existing technology, and the social values affected.
Several years is almost certainly the shortest period in
which a major innovation can fill a market opportunity.
2-5
-------
While the process of innovation can be shaped and directed
to some extent by legislation, research and development, and investment,
there are limits to how much it can be affected.
The issue is not only whether the evaluation period can be
shortened, but also --in the light of the long-term changes in the offing,
as noted in £, and the reductions in emissions from spark ignition engines
currently being achieved -- whether the evaluation period should be
shortened.
2.2 AUTOMOBILE SYSTEMS AND OPERATIONAL
REQUIREMENTS: INTERRELATIONSHIPS
INFLUENCING PERFORMANCE AND FUEL
ECONOMY*
The performance and fuel economy of an automobile is deter-
mined by detailed engineering characteristics of the specific vehicle, vehicle
operating requirements, and demands of the driver. The following provides
a. brief explanation of some of the basic interrelationahips of design and
operating parameters, providing a generalized background for the interpre-
tation of detailed evaluations of changes in vehicle design characteristics.
The basic physical law governing the motion of an automobile
is Newton's second law of motion. In order to place an automobile in motion
a force must be exerted through the vehicle driving wheels. The amount of
the force depends on the weight of the car and the rate of vehicle accelera-
tion, as well as the magnitude of forces which retard motion such as rolling
resistance and aerodynamic drag.
The motive force is applied to the vehicle driving wheels by
a combination of the operation of the vehicle engine, transmission, and
rear axle. The function of the transmission and rear axle is to deliver the
turning effort of the engine to the driving wheels at a speed different from
This section based principally on selected abstracts from Ref. 2-3.
2-6
-------
that of the engine. This difference in speed also implies a multiplication
of the turning effort of the engine, depending on the value of the overall gear
ratio.
The turning effort of the engine is termed torque and arises
as a result of ignition and subsequent expansion of the products of combus-
tion of a hydrocarbon fuel and air mixture within the cylinders of an internal
combustion engine. Detail design characteristics of specific internal com-
bustion engines determine the amount of torque produced by that engine.
Basically, however, it is the quantity of fuel and air ingested by the engine
during each revolution that determines the amount of torque produced.
Typically, small-displacement engines produce less torque than larger-
displacement engines, all other factors being equal.
Traffic conditions require that a vehicle be operated at varied
speeds and accelerations, all of which require varied torques. Ignoring the
transmission momentarily, the torque needed to sustain the desired operat-
ing condition is obtained by varying the amount of fuel air mixture processed
by the engine. For a vehicle with a carbureted engine this is, in effect.
accomplished by the modulation of the accelerator pedal. Depressing the
accelerator allows more mixture to enter the engine, thus producing more
torque. Of course, there is a limit to the amount of torque that any engine
with fixed design characteristics can produce. This maximum torque at any
given engine speed is produced at wide open throttle (WOT). The generalized
torque characteristic of an internal combustion engine is shown in Figure 2-1.
The transmission is used to multiply the torque output of
the engine so that the vehicle can be subjected to motive forces higher than
could be developed by the direct connection of the engine to the rear axle.
Transmission gear ratios greater than 1:1 allow the torque necessary to
move the vehicle to be delivered at smaller throttle (accelerator) settings.
To meet any given vehicle torque requirement, many gear ratios and various
combinations of engine displacement and throttle openings can be used.
2-7
-------
t
u
O
ce
O
l-
WOT
ENGINE SPEED
Figure 2-1. Generalized Engine Torque
Characteristic
Another basic concept which must be considered in the
discussion of automobile performance and economy is power. To define
what is meant by power we must go back to our explanation of the force
acting at the wheel of the vehicle. As the force acts on the vehicle, the
vehicle begins to move. The force, of course, then acts on the vehicle over
the distance through which the vehicle moves. This quantity (force X dis-
tance) is termed work, and the rate at which the work is done is termed
power.
In engineering terms, the power output of an engine is
specified as Power = Torque X rpm Constant. The constant is selected to
get the units of power into a comparative unit of measure, such as horse-
power. In essence, power is the price one has to pay when demanding a
given force at a given speed. It is the power requirement of the desired
motion of an automobile, relative to the overall efficiency of the translation
of the energy of combustion into meeting this power demand, that determines
automobile economy.
Z-8
-------
Before further deliberations on the fuel economy of
automobiles, it is necessary to note that other physical laws are also of
primary significance to the subject of fuel economy. The field of thermo-
dynamics provides guidance in this area and establishes theoretical effi-
ciency boundaries below which real systems can operate. In particular, no
practical internal combustion engine can have a thermal efficiency (power
output/power available from fuel combustion) over about 50 percent; i.e., it
is not possible to use more than 50 percent (and usually considerably less) of
the energy available in the fuel. The remainder is rejected to engine
exhaust and cooling systems. Detail design characteristics and operation
of various engines dictate how close to this theoretical limit real engines
can operate. An example of estimates of thermal efficiency for various
engines powering an automobile over the Federal Driving Cycle is given in
Figure 2-2. Note that the efficiency of a number of the alternative engines
equals or exceeds that for the spark ignition engine.
The conventional naturally aspirated (NA) carbureted engine
demonstrates a performance (torque vs rpm) map under fully warmed-up
conditions similar to Figure 2-3. The contours are lines of constant fuel
consumption for each unit of horsepower developed (brake specific fuel
consumption). This characteristic map results from two primary effects.
The first is the increase in brake specific fuel consumption (BSFC) with
decreasing torque output at a constant engine speed. This is due to the
higher percentage of friction power at low brake power outputs. Second,
brake specific fuel consumption increases with increasing speed (at constant
torque) due to increased engine friction. Consequently, the highest fuel use
efficiency for the engine occurs at low engine speeds and high torque outputs.
The engine performance map is an extremely significant
constituent of the overall vehicle fuel economy formulation. This map is
the representation of the units of fuel consumed in meeting the power
demanded by the motion of the vehicle. The values of the brake specific
fuel consumption, as well as the shape of the characteristic fuel-consumption
curves are extremely significant in determining accurate estimates of
economy.
2-9
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20r-
ro
i
UJ
o
I '5
UJ
a:
UJ
a
U 10
UJ
U
u.
u.
UJ
2
a:
UJ
DIESEL
(Mercedes
Benz)
MAX.
DUAL STRAT-
CHAMBER IFIED
STRATIFIED CHARGE
CHARGE (Texaco)
WANKEL(Honda
-(Mazda)
1973
ICE
MIN.
STIRLING
(Ford-
STRAT.F.EDPhllipS)
RANKINE CHARGE
BRAYTON (SES) (Ford)
(Chrysler)
Figure 2-2. Thermal Efficiency over the Federal Driving Cycle (Ref. 2-4)
-------
140
120
100
V)
Q.
UJ
2
ffi
80
60
40
20
HORSEPOWER PER SQUARE INCH OF PISTON AREA
0.6 1.0 1.4 2.0
I I
I I I
r i
1000 2000
PISTON SPEED - fpm
3000
1000
2000 3000
ENGINE rpm
(3. 5 in. stroke)
4000
5000
Figure 2-3. Performance Map of a Carbureted Spark
Ignition Engine (Ref. 2-5)
2-11
-------
Unfortunately conventional applications of engines to vehicles
result in a compromise of where the engine operates in the map under road-
cruising conditions relative to the maximum performance capability of the
vehicle. This is illustrated in Figure 2-4. The road load is met at points
of high BSFC, and the difference between the torque capability at full throttle
and the road-load torque value determines the performance (acceleration)
capability of the vehicle. If design characteristics of a vehicle were adjusted
such that the road-load power requirements could be met by operating near
the point of minimum BSFC, a marked improvement in fuel economy could
then be gained. However, such a move with conventional technology (such
as a very small engine) would also result in substantial reduction in the
acceleration and passing capabilities of a full-size automobile.
t
UJ
D
O
Of.
O
-ACCESSORY TORQUE
MAXIMUM TORQUE (WOT)
ROAD LOAD TORQUE
(Aerodynamic and rolling
resistance)
ENGINE SPEED
ROAD SPEED
Figure 2-4. Generalized Engine and Vehicle Torque
Characteristics
2-12
-------
The foregoing discussion stressed the difference between
the full power capability of a vehicle and the power necessary to maintain
level-road motion (no acceleration). For a full-size automobile, typical
road-load power requirements are given in Table 2-1.
Table 2-1. Typical Load-Power Requirements
(Ref. 2-3)
Speed (mph)
20
30
40
50
60
70
Power (hp)
5.0
8.2
13.0
19. 0
28.0
38. 5
This power requirement is determined by a large number of
factors including vehicle weight, tire design and construction, road surface
conditions, air temperature and density, and vehicle size and shape. The
vehicle design must accommodate a range of power requirements that result
from operation throughout the United States. For example, at an altitude
of 4,000feet, the road load at 70 mph is reduced by 3. 8 horsepower, or
about 10 percent, as a result of the decreased air density. Operating a
vehicle in a 0°F ambient instead of a 100°F ambient increases the road load
by about 12 percent at 70 mph. Fuel economy, then, can be seen to be
influenced by ambient conditions as well as the vehicle operating
r equi r em ent s.
The fuel economy measured under level-road conditions is
one method for comparing vehicle designs; however, the vehicle must be
capable of meeting varied road conditions, such as grades. "Level" sections
of interstate highways aren't always level. Often highways include grades,
2-13
-------
and a 3-foot rise in 100 feet of road is not uncommon. The conventional
American automobile can negotiate such grades without requiring downshift
to a lower gear. The significance of grades on vehicle-power requirements
is illustrated by Table 2-2.
Table 2-2. Three-Percent Grade Power Requirements
for a 4,300-pound Vehicle (Ref. 2-3)
Speed (mph)
20
30
40
50
60
70
Additional Grade
Horsepower
6.88
10. 32
13.75
17. 19
20.63
24.07
% Increase Over
Level Road
137.0
143.0
105.0
90.0
74.0
63.0
Negotiating a grade of only 3 percent more than doubles the
road-load power requirements at speeds up to 40 mph and also provides a
substantial increase even at 70 mph.
A vehicle can't be designed to maximize economy under
specific level road conditions but must be designed to accommodate a wide
range of driving conditions encountered in the United States. In addition
to steady-state driving conditions, transient operation of vehicles is also
extremely important. Stop-and-go traffic conditions such as accelerations,
braking, idling, etc. , constitute typical driving conditions experienced by
the driving public.
Under acceleration, weight is the single most important
parameter influencing vehicle-fuel consumption. Figure 2-5 clearly estab-
lishes the magnitude of this power demand for cars powered by spark
ignition engines. Under the maximum acceleration requirements of the
2-14
-------
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25
o
12°
s_
s.
V)
^ 15
E
o
o
o
UJ
UJ
D
U.
10-
1000
^AVG. 1957-1967
^ CARS
~ ' AVG. 1973 CARS
2000
3000 4000
INERTIA WEIGHT (pounds)
5000
6000
Figure 2-5. Fuel Economy over the Federal Driving Cycle (Ref. 2-4)
-------
EPA Urban Dynamometer Driving Schedule (also called the LA-4
dynamometer cycle), the power required to accelerate a 4300-pound vehicle
is over six times that required to maintain steady motion at 30 mph. (The
influence of vehicle weight can be marked; however, the significance over
a driving cycle will depend on the amount of time spent in acceleration
modes and the amount of acceleration demand.) Of concern to the motoring
public and the engine designer is the noticeable drop in fuel economy of the
later model cars that come equipped with more elaborate emission control
equipment.
For further comparison, Figure 2-6 illustrates the power
requirements of a 4300-pound vehicle under the same acceleration loads
and the power requirements of accelerating the vehicle at other performance
levels. It should be noted that the maximum performance level of a. full-
size automobile lies somewhere between 5 and 6 seconds to 30 mph. The
performance capability of a given vehicle weight then can be seen to signifi-
cantly influence the power required and, consequently, the fuel consumed.
The weight of the vehicle is the primary factor that deter-
mines the installed power requirements of the vehicle when a given level of
performance is specified. For example, in determining the power necessary
to accelerate a 4300-pound vehicle to 30 mph in 5 seconds, less than a
10 percent error would be encountered in totally ignoring the road load.
Conversely, unless the performance level or weight of a full-size vehicle is
significantly reduced, the maximum power output of the engine must remain
at previously used levels.
The influence on vehicle operating fuel economy due to starting
and operating a vehicle "cold" as compared to "warm" is significant. The
loss in fuel economy is attributable to varied sources including carburetion
enrichment to maintain the air fuel ratio of the engine cylinders, viscous
losses in cold engines, transmissions, and axles and higher heat losses in
the engine due to rejection of heat to "cold" coolant. Since these effects are
thermal, ambient temperature would be expected to influence the amount of
2-16
-------
100
QL
UJ
I
£ 50
WJ
a
i
I I
i r
ACCELERATION POWER (0-30 mph in 5 sec)(0.273 g)
ACCELERATION POWER (0-30 mph in 6 sec)(0.228 g
ACCELERATION POWER
LA-4 MAXIMUM ACCELERATION
(3.5 mph/sec)
ROAD LOAD POWER
REQUIREMENTS-
.—\ 1 i r
10
20
30
VEHICLE SPEED, mph
Figure 2-6. Motive Power Requirements, 4300-Pound Vehicle (Ref. 2-3)
-------
fuel economy degradation. Figure 2-7 illustrates this effect. These curves
were developed from sequential repetition of a driving cycle. Obviously,
the driver who makes short trips in cold climates will suffer a significant
reduction in his fuel economy from that obtainable from a fully warmed-up
engine.
Table 2-3 illustrates that over half of the trips taken in the
U.S. are under 5 miles. Although the total miles driven for these short
trips are not a major portion of the total travel, improvements in the fuel
consumption during warmup will also decrease the overall energy demand.
Table 2-3. Auto Trip Statistics (Ref. 2-6)
Urban Mileage
Trip Length
(One-Way Miles)
Under 5
5-9
10-15
16-20
21-30
31-40
41-50
51-99
100 and over
Total
Trips, %
54. 1
19.6
13.8
4. 3
4.0
1.6
0. 8
1.0
0. 8
100.0
Vehicle Miles, %
11. 1
13.8
18.7
9.1
11.8
6.6
4.3
7.6
17.0
100.0
Total Mileage, %
Urban 55.5
Highway 44. 5
2-18
-------
90IF AMBIENT TEMPERATURE
70 F AMBIENT TEMPERATURE
30 F AMBIENT TEMPERATURE
10 F AMBIENT TEMPERATURE
I I I
ill
10
TRIP LENGTH - MILES
15
Figure 2-7. Warmup Economy (Ref. 2-7)
-------
2.3 EXHAUST EMISSIONS AND EMISSION
CONTROL SYSTEMS
Gaseous emissions from automobile exhausts may be
controlled either by inhibiting formation of the gases in the engine cylinders
or by lowering their concentration externally. In general, methods of
controlling exhaust emissions from automotive spark ignition internal com-
bustion engines to meet the stringent 1975-76-77 Federal standards involve
certain engine modifications and the use of a combination of several devices.
Multiple methods are necessary because of the requirement for simultaneous
control of the hydrocarbon (HC), carbon monoxide (CO), and oxides of
nitrogen (NO ) constituents in the exhaust gas.
3C
2.3.1 Engine Modifications and Operating Considerations
As illustrated in Figure 2-8, HC, CO, and NO concentrations
Ji.
in the exhaust of uncontrolled engines are strongly a function of the operating
air fuel ratio. As can be noted from the figure, at the stoichiometric air
fuel ratio NOx production is very high while HC and CO production is rela-
tively low. For air fuel ratios between approximately 17 and 19, levels for
all three constituents are reduced considerably from peak values. Currently,
engine operation is restricted to air fuel ratios below approximately 19 to
avoid excessive power loss and rough engine operation. Operation in the
17 to 19 range minimizes HC and CO levels, and lowers that of NO , but the
x'
concurrent reduction of exhaust levels of all three species is far from suffi-
cient to meet 1975-76 emission requirements. NO formation can be
suppressed by operating in the "rich" regime (air fuel ratios of approxi-
mately 11-13); however, in this region HC and CO concentrations are very
high.
Other factors affecting emissions include spark timing and
design of the induction system and combustion chamber. Retarding the
spark results in lower peak temperatures and less NOx formation. Also,
the exhaust gas temperature is higher with a retarded spark, which promotes
2-20
-------
o —
8
o
LU
UJ
LJ
QL
STOICHIOMETRIC
AIR FUEL RATIO
14 16 18
AIR FUEL RATIO
Figure 2-8. Effect of Air Fuel Ratio on Emission Levels,
Gasoline Spark Ignition Engine (Ref. 2-8)
2-21
-------
further combustion of the HC and CO species in the exhaust system.
Induction-system modifications can result in lower emissions by providing
a more uniform mixture to the cylinders and better atomization and vapori-
zation of the fuel. Combustion chamber design affects the combustion
process and, as a result, peak and exhaust gas temperatures.
2.3.2 Applicable Passenger Car Emission Standards
Table 2-4 summarizes the applicable statutory emission
standards for gasoline-powered passenger cars as currently promulgated.
As can be noted, the Federal standards apply nationwide except for those
instances where California has been granted a waiver to adopt more strin-
gent control of one or more emission species in a given time period.
Through model year 1974 these standards have been met by
the combination of engine modifications (combustion chamber redesign,
spark retardation, etc.) and the incorporation of exhaust gas recirculation
(EGR) systems for control of NO to meet 1973-74 standards.
x
2.3.3 Projected Emission Control Systems
It is generally agreed that the emission control system
required to meet the California 1975 standards (see Table 2-4) is exemplified
by the following package of components and engine modifications:
Oxidation catalytic converter
Air injection
Exhaust gas recirculation (EGR)
Carburetor modifications
Ignition system modifications.
There is not complete agreement as to whether the catalytic converter will
be required to meet the less stringent Federal 1975 standards that apply to
the rest of the nation. However, some manufacturers have implied that at
least some of their car models sold in states other than California will also
2-22
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Table 2-4. Summary of Passenger Car Emission Standards
(Gasoline Powered)
1
OO
Applicable Year and Area
Test Procedure
Hydrocarbons
Carbon Monoxide
Emission
Nitrogen Oxides
Evaporative
(grams per test)
Emission Rate, Grams/Mile
Uncontrolled
Pre-1966a
FTP CVS-C
10 17
77 125
4-6 4-6
40
1966 Cal
1968 Fed
1970
Cal Fed
1971
Cal Fed
FTP (Concentration Measure)
3.4
35
NR
NR
2.2 2.2
23 23
NR NR
6 NR
2.2 2.2
23 23
4 NR
6
1972
Cal Fed
1973
Cal Fed
1974
Cal Fed
CVS-C
3.2 3.4
39 39
3.2b NR
2
3.2 3.4
39 39
3.0 3.0
2
3.2 3.4
39 39
2.0 3.0
2
1975°
Cal Fed
1976d
Nation
1977d
Nation
CVS-CH
0.9 1.5
9.0 15
2.0 3.1
2
0.41
3.4
2.0
2
0.41
3.4
0.4
2
FTP 7-mode Federal test procedure
NR No i
CVS-C CVS
CVS-CH CVS
equirement
cold/hot weighting mass test
3Estimate only, for comparison purposes
Two hot cycles of FTP test
COriginal 1975 standards were: HC = 0.41, CO = 3.4, and NOx = 3. 1
Original 1976 standards were same ae current 1977 standards
-------
incorporate the catalytic converter. An example of emission control system
elements for control of both exhaust and evaporative emissions is given in
Figure 2-9 for General Motors' projected 1975 design.
The oxidation catalyst with air injection is provided to pro-
mote the oxidation of the unburned hydrocarbon (HC) and carbon monoxide
(CO) species contained in the engine exhaust. The catalyst type which
appears most frequently among the selected first-choice systems is the
platinum-group metal/monolithic converter. Base metal/pelletized and
promoted base metal/pelletized catalyst designs* also are being evaluated
by some manufacturers. Some automobile manufacturers are also con-
sidering the utilization of a catalyst overtemperature protection device to
prevent catalyst damage under extreme or abnormal engine operating condi-
tions (spark plug misfire, etc. ).
Exhaust gas recirculation (EGR) systems for suppression of
NO formation will be employed in nearly all of the 1975 model year auto-
5C
mobiles. These systems will be improved versions of the EGR systems
used in most of the 1973 and 1974 model year vehicles.
The emission control systems of a number of manufacturers
utilize air injection in the exhaust manifold to provide more rapid warmup
of the catalyst under cold-start conditions. A full-size thermal reactor is
expected to be used in the 1975 system for the Mazda rotary engine.
Carburetion/intake system modifications range from complete
redesigns utilizing new concepts to minor improvements to the current con-
ventional systems. These modifications are generally directed toward im-
proving the precision and stability of the air fuel ratio and also include such
features as altitude compensation, quick-release choke devices, and intake-
manifold heating.
* Promoted base-metal catalyst formulations contain small amounts
of platinum-group metals.
2-24
-------
-AIR INJECTION
PUMP
-QUICK HEAT
MANIFOLD (EFE)
IMPROVED CARBURETION AND CHOKE
ALTITUDE AND TEMPERATURE
COMPENSATION
EXHAUST GAS
RECIRCULATION
MODIFIED SPARK
TIMING
I
IV
CATALYTIC
CONVERTER
DOMED TANK
VAPOR SEPARATOR
CARBON
CANISTER
-ELECTRONIC
IGNITION
Figure 2-9 . General Motors Projected 1975 Under-Floor Emission
Control System (Ref. 2-9)
-------
All domestic manufacturers propose, or have in development,
electronic (breakerless) ignition systems which are targeted for inclusion
in their first-choice emission control system. These systems generally
provide an improvement in spark-timing precision, consistency, and
reliability.
Emission control systems for meeting 1976 standards are
projected to be similar to the 1975 systems just described, with improve-
ments in catalytic converters and associated components to meet the lower
HC and CO requirements.
Emission control systems that have been under consideration
by the automobile manufacturers to meet the stringent NO standard pre-
3C
scribed for 1977 include all components of the 1976 system plus:
Reduction catalyst(s) installed upstream of the
oxidation catalyst(s)
More sophisticated air injection systems
Modified carburetion, ignition, and EGR systems.
To date, no manufacturer has reported acceptable durability for reduction
catalysts, however further development testing is in progress.
2-26
-------
SECTION 3
-------
3. GAS TURBINE ENGINES
3. 1 INTRODUCTION
3. 1. 1 General Description
In its simplest form, the gas turbine engine consists of a
compressor, a combustion chamber, and a turbine (see Figure 3-1). Air is
taken in by the compressor at atmospheric pressure, compressed to a higher
pressure, and is then delivered to the combustion chamber where fuel is in-
jected and burned at essentially constant pressure. The resulting high-
temperature, high-pressure gas is expanded in the turbine and then exhausted
to the atmosphere. Part of the turbine-shaft work is used to drive the com-
pressor; the remainder provides useful output work.
Several modifications of this simple cycle have been utilized
in the development of gas turbine engines for vehicular use. Most engines
currently in use for on-highway vehicles employ a free power turbine; i.e.,
two mechanically independent turbine stages are used (see Figure 3-2).
One drives the compressor and the other drives the output shaft. This allows
a wide range in output-shaft speed, which is not possible with a single-
shaft turbine engine. To improve efficiency, such free turbine engines use
a regenerative cycle. In a regenerative cycle, the energy in the turbine ex-
haust is utilized to raise the temperature of the compressor discharge flow,
thereby reducing the energy input requirement for the combustor (Figure 3-3).
The heat transfer is accomplished by means of a regenerator or recuperator.
A regenerator is a rotating-disc heat exchanger that alternately exposes the
heat exchanger core by sections to turbine exhaust gases and compressor out-
let air. A recuperator is a fixed (static) heat exchanger with separate pas-
sages for turbine exhaust gases and compressor outlet air.
3. 1.2 Historical Background
The fundamental principles of the gas turbine engine were well
known by the end of the eighteenth century; however; turbomachinery operating
3-1
-------
FUEL IN
AIR INTAKE
X
r
OUTP
COMPRESSOR
TURBINE
Figure 3-1. Simple Gas Turbine Engine, Schematic Diagram
FUEL IN
\
COMF
\
r^1*
COMBUSTOR
>RESSOR COMP
RESS
1
OR
^
x
OUTPU
TURBINE POWER
TURBINE
GAS-GENERATOR SECTION
Figure 3-2. Free-Turbine Engine, Schematic Diagram
3-2
-------
EXHAUST
COMBUSTOR
AIR
INTAKE
i
OO
n/W <
REGENERATOR
COMPRESSOR
COMPRESSOR
TURBINE
POWER
TURBINE
-OUTPUT
a SHAFT
GAS-GENERATOR SECTION
Figure 3-3. Regenerative Free-Turbine Engine,
Schematic Diagram
-------
successfully on the gas turbine thermodynamic cycle, called the Brayton
cycle, is a fairly recent development. The first really successful power
plant was the aircraft turbojet developed from intensive work begun in the
1930s. The development of gas turbine engines for automotive use was begun
immediately after World War 11 by Chrysler in this country and Rover in
Great Britain. The first Rover turbine-powered automobile ran in 1959.
General Motors began automotive gas turbine work around 1948 and Ford
about 1950. These developers selected the free turbine configuration early
and moved quickly to the regenerative configuration.
Principal early automobile turbine problems were noise,
poor fuel economy, poor durability, acceleration lag, and high manufactur-
ing cost. Considerable improvements have been made over the years so
that most of these items do not constitute serious deficiencies at the present
time. It is expected that the next generation of gas turbine engines will be
further improved, particularly in the area of fuel economy. The recent
awareness of domestic energy shortages has placed additional emphasis on
improving fuel economy in the gas turbine program sponsored by the Alter-
native Automotive Power Systems (AAPS) Division of EPA. In the far term,
if ceramic turbines are developed, a substantial increase in turbine inlet
temperature would be possible, leading to a significant reduction in fuel
consumption.
3.2 POWER PLANT DESCRIPTION
3. 2. 1 Power Plant Configuration
Most existing vehicular gas turbine engines use the so-called
free turbine arrangement wherein two turbines are used in series. In this
arrangement, a high-pressure, first-stage turbine* drives the air compres-
sor and a low-pressure, second-stage turbine' ' delivers useful shaft power
Also referred to as the gasifier or compressor turbine
£
Also referred to as the power turbine
3-4
-------
output. This split-shaft scheme allows for flexibility of operation,
expecially in conjunction with a variable power turbine nozzle. For exam-
ple, the variable turbine nozzle may be used to provide engine braking in
addition to its value in providing high torque at low turbine speed.
The single-shaft arrangement in which the turbine supplies
both the compressor power and the net power output has been used exten-
sively in electric power generation applications where the gas turbine oper-
ates at a fixed design speed. The torque/speed characteristics of the single-
shaft engine are very steep and the torque is virtually not available below
about 50 percent peak speed. Therefore, the configuration requires a wide-
range, multiple step or a continuously variable gear-ratio transmission.
A simple gas turbine cycle utilizes a compressor which in-
creases the air pressure, a combustor in which fuel is added and burned
with the air, and a turbine powered by the hot combustion products. The tur-
bine drives the compressor and also produces useful work. A temperature-
entropy diagram depicting this process is shown in Figure 3-4. The thermal
energy in the simple cycle turbine exhaust is wasted. In a regenerative gas
turbine cycle, the energy in the turbine exhaust is used to raise the tempera-
ture of the compressor-discharge flow which, for a given engine design
power level, reduces the energy input requirement for the combustor. The
heat transfer is accomplished by means of a regenerator or recuperator.
The regenerator is a rotating-disc heat exchanger that alternatively exposes
the heat exchanger core by sections to turbine exhaust gases and compressor
outlet air. The recuperator is a fixed (static) heat exchanger with separate
passages for turbine exhaust gases and compressor outlet air. A schematic
of this equipment is presented in Figure 3-5.
3.2.2 Design Features
No automobile turbine engine is in production today, although
several experimental gas turbine engines have been developed by automobile
and gas turbine engine manufacturers. Some of these engines and their
design features are shown in Table 3-1. In addition to these experimental
3-5
-------
UJ
QL
ID
tt
UI
O.
2
LJ
h-
TURBINE
CONSTANT PRESSURE
ACTUAL PROCESS
ENTROPY
Figure 3-4. Temperature-Entropy Diagram for Simple
Cycle Gas Turbine
3-6
-------
OJ
I
Hx2
REGENERATOR
HEAT EXCHANGER
Hxl
Cx2
't2
Figur-e 3-5. Equipment for Gas Turbine Regenerative Cycle,
Schematic Diagram
-------
Table 3-1. Engine/Vehicle Basic Characteristics
Manufacturer
Ford
CM
AiResea rch
United
Airc raft
Wilh&mi
Reiea rch
Chrysler
Rover
Fiat
Volvo
Boeing
GE
Solar
Type of Engine
Three Shaft with
Intercoolmg. Re-
heat, Recuperator
Regenerative Cycle
with Ceramic
Regenerator
Free Turbine
with Metal Regen-
erator
Single Shaft Regen-
erative Cycle with
Metal Regenerator
Single Shaft Recu-
perative Cycle with
Ceramic
Recuperator
Free Turbine Re-
with Ceramic
Regene rator
Simple Cycle
Single Shaft
Simple Cycle
Free Turbine
with Regenerative
Single Shaft
*e"eratlve *
Free Tu r b i n e
Regenerative Cycle
with Ceramic
Regenerator
Regenerative Cycle
Free Turbine
Regenerative Cycle
with Regenerative
Free Turbine
Recuperative Cycle
with Metal
Recuperator
Simple Cycle
Free Turbine
Single Shaft
8 V
Simple Cycle
Free Turbine
with Ceramic
Regenerator
Recuperative Cycle
with Ceramic
Recuperator
Recuperative Cycle
with Metal Recu-
perator Free
Turbine
hP
3ZO
450
525
225
225
280
325
.551
125 '
I751
150 '
ISO1
.63'
166 '
80
13G1
150
150
200
250
400
150 1
150'
300
Peak
Operating
Speed, rpm
46, 500/91, 500/
37. 500 z
45, 300/38, ZOO3
37.500/31,650
37. 500/31.650
44.000/ ?
33.000/27,000
37.100/ 30.830
70.230
83,600
67. 500/55.200
130,000
130,000/40.000
1 12.000
66. 500/71. 300
61.000/51,000
44.600/45,700
65,000/36,000
29,000/22,000
43.000
40.000
80.000/50.000
36.400/31.000
Dimension*, in.
Length
40.8
7
37
36
47
17
21
26
27. 5
24
33
38
51.2
39
34.5
31. 5
50
Width
34.6
?
26.4
30
28
24
27. 5
30
26.4
27
26.5
27.0
36
Height
41.8
7
24
36
42
19
17
21
28
16
29.5
27
29.5
24
24. 1
24. 5
41
Specific
Ib/hp
S. 31
3.78
3.23
7
2.68
3.4
5.23
2.2
3.2
2. 5
1.07
1.33
1.78
2.95
3. 13
1.4
4.42
3. 17
2.90
3.26
1.0
4.79
4.57
5. 1
Deiign SFC
Ib Fuel/
hp Hr
0.41
0.41
0.43
0.551*
0.54
0.45
0.475
0.41
0.43
0.42
0.51
0.51
0.45
0.53
0.59
0.40
0.60
0. 55
0.95
0.40
0.75
0.65
0. 50
0.60
i'r* ^ - Engine, not built or under test
"Three speeds shown for low pressure spool, high- pressure spool, and free turbine, respectively
foT° "wVrd" '^ *M f°r du*l"Sh*fl (free turbin*> engines; first figure for first stage turbine, second figure
*r-
f-xperimert.il -.tH bed engine.
3-8
-------
engines, the EPA-sponsored gas turbine studies resulted in several engine
designs for a nominal 4,000-pound vehicle. The design features of these
engines will be discussed in the following sections.
3.2.2.1 Power
Design power levels for several experimental engines and the
EPA-sponsored study engines are shown in Table 3-1. They range from an
80-hp engine for an AMC Hornet designed by Williams Research to a 525-hp
truck engine built by Ford. The range of power for the paper-study engines
was from 125 to 175 horsepower.
3. 2. 2. 2 Engine Size and Weight
Size and weight of the engine impact the power train packag-
ing, frame, suspension, and other vehicle design elements. In general, both
specific weight and specific volume of gas turbine engines are lower than for
equivalent spark ignition engines. Weight and volume data for a number of
gas turbine engines are provided in Table 3-1. The specific weight, and spe-
cific volume values for a family of engines, when plotted against horsepower,
maybe expected to decrease as the horsepower increases. However, varia-
tions in the data shown in Table 3-1 do not permit such a trend to be delinea-
ted. Note that the simple-cycle engines are significantly lighter than the
regenerative engines.
The dimensional data available do not permit an accurate cal-
culation of engine volume; therefore, only overall engine dimensions are
shown in Table 3-1. The simple-cycle engines are also seen to be relatively
compact as compared to the regenerative engines. In each category the
engine size appears to be inversely related to peak operating speed.
3.2.2.3 Engine Components
The efficiencies of compressor and turbine have direct effect
on cycle efficiency and the net work output of the system. An efficient com-
bustor is important from the standpoint of emission and fuel economy con-
siderations. Careful design of regenerators and recuperators in a regener-
ative cycle is important to ensure high heat exchanger effectiveness and low
3-9
-------
pressure losses. Other components, such as bearings, have a significant
impact on maintaining mechanical integrity of the engine. Engine control
components also have to be defined to ensure that proper response and per-
formance designed into the engine are realized over the complete operating
map. These components are discussed in the following sections.
3.2.2.3.1 Compressors
The compressor design may be either axial or centrifugal.
In large gas turbines of the type used for jet aircraft or stationary power
equipment, axial compressors are typically used because of their superior
efficiency at the high flow rates of these applications. In an automotive gas
turbine, where the maximum airflow rate is approximately 2. 5 Ib/sec, the
blade height of an axial compressor is very small, and maintaining good
efficiency is a problem. Hence, in this application, the centrifugal com-
pressor is preferred because, in addition to eliminating the blade-height
problem: (1) the centrifugal impeller (wheel) is rugged and easier to man-
ufacture, (2) maintenance problems are fewer, (3) the axial length is
shorter (which improves engine packaging and bearing-support loads), and
(4) the inertia is lower (which improves engine response).
Important parameters affecting the performance of centrifugal
compressors are the clearance between the impeller and the shroud and the
ratio of impeller-to-shroud clearance to blade width at the tip of the impeller.
In general, the relationship between efficiency loss and tip clearance is
linear (Ref. 3-2). Typically, the clearance losses vary between less than
1 percent for large compressors to about 3 percent for very small units.
With regard to impeller materials, aluminum can be used for
pressure ratios up to about 4. 5:1. Beyond that point, steel alloys and
titanium are required in single-stage designs.
3.2.2.3.2 Turbines
Like compressors, either axial or radial turbine configura-
tions may be employed in gas turbines. In large gas turbine engines, the flow
rates are high and the turbines used are axial. For automobile gas turbines
3-10
-------
with maximum flow rates of approximately 2.5 Ib/sec, axial or radial
turbines are feasible. Radial turbines are more rugged than the axial type,
and their performance is less sensitive to variations in tip clearance. They
can also be operated at higher speed. In single-shaft arrangements, a radial
turbine is preferred. However, for split-shaft designs in which the two tur-
bines are in series, axial turbines are preferred because of packaging con-
siderations for the routing of gas-flow passages. A typical turbine perfor-
mance map is presented in Figure 3-6 showing the relationships between
corrected flow, pressure ratio, and corrected speed.
o
LU
I-
o
LU
QL
Of.
O
O
0.40
0.35
0.30
0.25
0.20
0.15
0.10
0.05
160 N//0" (corrected speed)
3456789
COMPRESSOR PRESSURE RATIO
10 11
Figure 3-6. Estimated Turbine Performance, Single-
Shaft Recuperated Engine (Ref. 3-1)
3-11
-------
A critical dimensional tolerance impacting performance as
well as mass production quality control requirements is the tip clearance
for the axial turbine wheel and the shroud clearance for the radial in-flow
turbine.
State-of-the-art turbine design indicates the possibility of
nominal design point operation at 1,900°F turbine-inlet temperature for
super alloy investment castings without blade cooling. For higher temper-
atures, blade cooling or advanced ceramic materials would be required.
3.2.2.3.3 Combustors
In the past, when emission control was not a design objective,
the prime consideration in gas turbine combustor design was simply to pro-
duce the highest overall combustion efficiency in a compact configuration.
The basic approach now is to create a hot combustion zone by injecting fuel
in a recirculating flow region (primary combustion zone) where near-
stoichiometric air/fuel conditions are maintained. The high primary
zone gas temperatures are reduced in the secondary zone of the combustor
by mixing the combustion products with dilution air.
An example of a conventional combustor configuration is the
device selected by Ford for its 707 heavy-duty gas turbine engine. The design,
illustrated in Figure 3-7, is a reverse-flow, can-type combustor which
utilizes a fuel-atomizing nozzle and an air swirler installed in the dome of
the unit. In this design, a strong toroidal vortex is generated along the
combustor axis which, according to Ford, assures stable combustion at
all operating conditions of the engine. Approximately 20 to 30 percent of
the total engine air flow is utilized in the primary zone to provide a primary
air/fuel mixture ratio between about 20 and 35. The combustion gases are
diluted with secondary air injected downstream of the primary zone in such
a manner that a uniform flow velocity profile is obtained at the entrance to
the gasifier. This design satisfied Ford's initial objectives for stable com-
bustion, high combustion efficiency, low pressure drop, long life, low cost,
and low emissions.
3-12
-------
SWIRLER
VANES
PRIMARY
HOLES -
SECONDARY
HOLES
r
FLOW REVERSAL
GENERATED BY
FUEL NOZZLE
Figure 3-7. Flow Path for Reverse-Flow, Can-Type
Combustors (Ref. 3-3)
Present goals in automotive gas turbine combustor design
are to meet or better the original 1976 Federal emission standards while
maintaining good fuel economy. A major problem arises from the fact that
superior fuel economy requires regenerative heating and/or high compres-
sor pressure ratio, resulting in high combustor temperatures conducive to
high NO emissions. Figure 3-8 illustrates this relationship. As the
combustor-inlet temperature is increased (for the same equivalence ratio)
the combustor temperature and the NO levels increase. To minimize the
.X
formation of NO , the following design principles are currently being eval-
5x
uated by industry (Refs. 3-4 through 3-7):
a. Lean primary zone operation
b. Prevaporization of the fuel before combustion
c. Air/fuel ratio control by means of variable geometry
air ports.
3-13
-------
FLAME
TEMPERATURE
- °F
60001-
5000
4000
3000
2000
1000
V 20
1 1
30
AIR FUEL
1
40
RATIO
1
50
1
60
1.0 0.7
0. 5 0.4
EQUIVALENCE RATIO
EQUILIBRIUM NOX ,0000
CONCENTRATION - ppm
1000
100
10
0.3
cx2- '
1500
20
30 40
AIR FUEL RATIO
50
60
Figure 3-8. Estimated Flame Temperatures and Equilibrium NO
Concentrations as Functions of Air Fuel Ratio for X
Various Combustor Inlet Temperatures (Ref. 3-4)
An example of a fixed-geometry combustor configuration is
the current Chrysler combustor. The primary zone reaction volume is very
small, in order to minimize the residence time of the combustion gases at
the high combustion temperatures and, hence, to minimize the formation of
NO . Secondary air is injected immediately downstream of the primary
zone to quench the NO reactions and complete the hydrocarbon (HC) and
X.
carbon monoxide (CO) oxidation reactions.
3. 1. 2. 3. 4 Regenerators and Recuperators
Regenerators are rotating heat exchangers which operate by
alternately exposing the core sections to the "cold" and "hot" compressor
and turbine exhaust flows. As illustrated in Figure 3-9, a barrier seal
3-14
-------
HIGH-
PRESSURE
SIDE
LOW-
PRESSURE
SIDE
ML.
TURBINE
EXHAUST—*-
GAS
^
^
COMPRESSOR
OUTLET AIR
HEATING
ELEMENTS
Figure 3-9. Disc-Type Rotary Regenerator (Ref. 3-1)
separates the "cold" and "hot" sides of the heat exchangers in -which both
metallic and ceramic cores have been used. The effectiveness of these
rotating heat exchangers, defined as the ratio of heat transferred into the
cold fluid to the amount of heat available in the hot fluid, can be very high.
Values of the order of 0.9 have been achieved -with regenerators designed
for automotive gas turbine applications.
In the past, leakage from the high-pressure side to the low-
pressure side has been a major problem -with regenerators; however, this
now appears to be solvable.
Recuperators are fixed-boundary, stationary heat exchangers.
For automotive applications, the effectiveness of these heat exchangers can
be as high as 0.85. Compared to regenerators, recuperators may be bulky
but require no rubbing seals or rotating drive. Materials used in the manu-
facture of recuperators include mild steel, stainless steel, Inconel, and
Hastelloy (Ref. 3-8). Mild steel has been used for low-temperature
3-15
-------
applications in stationary gas turbine plants. An excellent recuperator
material is 347 stainless steel, which has good creep strength and hot ero-
sion resistance at elevated temperatures (Ref. 3-8). The cost of the high-
temperature superalloys has led to the investigation of ceramic materials
for both recuperator and regenerator applications.
From a packaging, performance, weight, and cost point of
view, the regenerators are preferred for automotive gas turbines. Cur-
rently, the glass-ceramic regenerator is growing in favor. The two major
suppliers of ceramic regenerator cores are Corning Glass Works, which
manufactures Cercor (Ref. 3-9). and Owens-Illinois, which makes Cer-Vit
(Ref. 3-10). Both manufacturers utilize honeycomb construction; the Cercor
product has straight triangular passages and Cer-Vit has hexagonal passages.
The advantages claimed for glass-ceramic regenerators are compactness,
low density, and low cost. The Cercor density runs between 30 to 42 Ib/ft .
The cost of a 17-inch diameter Cercor regenerator is projected to be about
$40 based on an annual production rate of 100,000 units.
The employment of such ceramic materials requires an under-
standing of their properties and limitations as well as the development of
satisfactory component design techniques. A major problem area involves
thermal stresses. Establishment of material and design requirements
necessitates a careful heat-transfer analysis of steady running and transient
operation to determine thermal stresses. For a regenerator application,
where the very nature of its function demands a high-temperature gradient,
glass ceramics of the lithium-alumina-silicate (LAS) type with their very
low thermal expansion are very attractive.
Another consideration in material selection is resistance to
chemical attack. Oxidation is the most obvious concern, but combined salt
ingestion in combination with small amounts of sulfur in the fuel creates at
high temperature a corrosion condition known as sulfidation. LAS is sus-
ceptible to such corrosion but intensive development is under way to provide
satisfactory protective coatings or substitute materials.
3-16
-------
Chrysler experimented with ceramic regenerators as early
as 1950, when it dropped the ceramic approach in a decision to use all-metal
regenerators in its fourth- and sixth-generation automobile gas turbine
engines. Chrysler feels that metal regenerators have lower seal friction and
are less expensive at current production levels than ceramic configurations.
In terms of heat loss, ceramic regenerators are considered superior. Even
though Chrysler continues to favor metallic regenerators for reasons of
mechanical performance, reliability, and cost, they are nevertheless working
on the development of improved ceramic designs.
3.2.2.3.5 Bearings
Both rolling element and journal bearings are used in gas
turbine engines. The rolling element bearing is preferred in aircraft gas
turbines because of superior friction characteristics under variable temper-
ature and load conditions. In stationary gas turbines, journal bearings are
used to take advantage of their long service life (Ref. 3-11). For automobile
gas turbines, both journal and rolling element bearings are being considered.
3.2.2.4 Materials
The production potential of gas turbine engines is dependent
on the types, cost, and availability of materials as received at the engine
plant and on the processes used to manufacture the engine. These factors
are reflected in production rates, equipment design, facilities, and the
eventual production cost of the engine. The following paragraphs discuss the
availability of raw materials and the requirements of processed materials
as well as their impact on supply industries, engine manufacturing processes,
equipment tooling and facility requirements, engine production costs, and
possible means for reducing production costs. Most of these items are com-
pared to current and future emission-controlled piston engines so as to
provide a basis for assessing effects incurred by a changeover to gas turbine
engine manufacture.
3-17
-------
The piston engine without advanced emission controls has
temperature/stress requirements that can be readily satisfied by the use of
low-alloy carbon steels, cast iron, and aluminum; piston engines with such
emission controls, as presently conceived, will require a certain amount of
high-temperature metal in the exhaust system. In contrast, the gas turbine
engine operates continuously at high temperatures and stresses and, there-
fore, requires a significant amount of materials having high-strength prop-
erties at elevated temperatures. These considerations govern the difference
in basic materials selected for the different engines.
Table 3-2 provides an approximate comparison of the types
and weights of materials used for contemporary and future piston engines
as well as for a low-pressure-ratio regenerative gas turbine; this comparison
is based on data provided in Reference 3-12. Data regarding the use of
Inconel 718 and 738 have been deleted from Table 3-2 because, by comparison
with Inconel 713, Inconel 718 contains much greater levels of niobium, and
Inconel 738 contains much greater levels of tungsten and tantalum. Most
manufacturers have elected not to use these potentially critical elements for
their engine designs. It should be noted also that there are considerable
variations in material weights for gas turbines proposed by a number of
manufacturers. However, compared to piston engines, all concepts use
relatively large amounts of high-alloy steel, Inconel, and Hastelloy.
Table 3-3 compares the 1969 U.S. consumption of potentially
critical materials with the material needs for production of 10 million engines
(the approximate quantity of passenger automobile engines produced annually
in the United States in recent years). These data were extracted from various
sections in References 3-1, and 3-12 through 3-15, and are based on initial
quantities of materials required with fabrication-scrap recycling. As can be
seen, some of the material requirements for the gas turbine engine, when
contrasted with 1969 U. S. consumption levels, would appear to present a sig-
nificant impact on future availability of these materials. However, some
lessening of the impact is anticipated. First, inquiries to stainless steel
producers have revealed that there has been a significant increase in
3-18
-------
Table 3-2. Engine Material Content and Weight (based on Ref. 3-12)
Material
Gray Cast Iron
High-Temp Cast Iron
Doctile Cast Iron
High Alloy Steel
Low Alloy and Carbon
Steel
Copper
Brass
Stainless Steel
ACI Type HU Steel
A luminum.
Inconel 713
Hastelloy-X
Cercor (A1-O-)
Platinum/ Palladium/
Ruthenium
Total
Ib/engine (scrap weight not included)
Piston Engine
Circa 1970
220
--
20
5
189
19
25
--
--
20
--
--
498
Circa 1976
(Est. )
224
28
20
5
226
19
25
5
25
21
--
--
0. 007
598
Gas Turbine
4:1 Regenerative
(Est.)
--
23
252
65
68
15
--
14
--
4
8
4
20
--
474
3-19
-------
Table 3-3. Estimated High-Temperature Metal Element Requirements
for 10 Million Engines* (Refs. 3-1, 3-1 2 through 3-1 5)
(Millions of Pounds)
I
M
O
Kl cmerit
Ch roniium
Nickel
Cobalt
Molybdenum
Tungsten
Niobium
Platinum Type
Annual Needs
Piston Engine
Circa 1 970
8
Circa 1976
147
141
0. 1
Gas Turbine
4:1 Regenerative
54
124
1.4
8. 5
2.4
1.6
Usage and Availability
United States
1969
Annual
Consumption
481
283
15
52
16
3
0. 1
1972
Stockpile
5331
79
74
43
129
9
0. 14
World
1970
Rese rves
1, 550, 000
150, 000
4, HI H
1 1, 000
2, KOO
13, 000
31
''Amounts contained in the engine after fabrication
-------
availability of chromium and nickel since 1969. Second, a further expansion
in production is expected in order to meet the requirements for the 1976 pis-
ton engine, unless some drastic changes in the present emission control con-
cepts occur. Therefore, cobalt, niobium, and molybdenum are the only
materials that may be critical in availability.
The above analysis for Table 3-3 considered the initial demand
for critical elements but did not include the recycling of scrapped engines to
extract the critical elements. It is estimated that, for the current annual
engine production rates, the quantities of high-temperature material elements
to be obtained from ore could be eventually reduced by a factor of 2 to 3 by
recycling all of the scrapped engines. Since the introduction of gas turbines
would probably build up from initial production quantities of approximately
250,000 engines per year to the full potential in excess of 10 million, the
material requirements in Table 3-3 eventually would not have to be obtained
from ore alone were recycling to prove profitable.
Aside from problems of availability, consideration must be
given to (1) potential effects on cost of the high-temperature elements due
to supply/demand factors, and (2) the balance-of-trade effect on the U.S.
economy due to the requirements of increased purchases of ore or refined
metals from foreign countries.
3. 2. 2. 5. Fuel Requirements
A gas turbine can use either liquid or gaseous fuel. Liquid
fuel requires atomization for efficient burning; the gaseous fuel does not
require any special preparation. Gas turbine engines can use gasoline,
kerosene, or diesel fuel. The same burner in the gas turbine engine can
accept several types of liquid fuel, possibly without adjustments for a fam-
ily of hydrocarbons. In using gaseous fuel, a different type of orifice sys-
tem is required.
3.2.2.6 Pressure Ratio and Temperature
To increase the net work output of the gas turbine, the com-
pressor work has to be reduced, the turbine work has to be increased, or
3-21
-------
both. An increase in the turbine inlet temperature has a significant effect
on the turbine work output. However, the operational turbine inlet tempera-
ture is limited by material properties. Present-day gas turbines can operate
at turbine-inlet temperatures of 1,7008F continuously and at temperatures
as high as 1,900°F for a short time. As will be discussed later, with ce-
ramic turbine materials this temperature can be increased at 2,500°F.
As the engine-pressure ratio is increased at constant tur-
bine inlet temperature, the temperature difference between the turbine and
compressor discharge decreases and regenerative heating becomes less
significant. For this reason, the design point efficiency of the simple cycle
optimizes at high pressure ratio, and the efficiency of the regenerative cycle
optimizes at low pressure ratio. This performance effect is shown in Fig-
ures 3-10 and 3-11, where the brake specific fuel consumption (BSFC) is
plotted as a function of compressor-pressure ratio for a range of turbine
inlet temperatures for typical simple and regenerative cycles, respectively.
However, the selection of pressure ratio in design studies will be influenced
by engine performance at part-load operation--where a major portion of
operating times is expected to be experienced by engines powering automo-
biles to be driven in urban traffic conditions.
In all cycles, the pressure ratio available in the turbine is
reduced by the combustor-pressure drop. In the regenerative cycle, addi-
tional pressure losses occur in the "cold" and "hot" flow passages of the
heat exchanger, which further reduces the turbine-pressure ratio and the
turbine work output.
3.2.2.7 Transmission Systems Design Approach
The gas turbine engine, as described in previous sections,
generally operates at speeds above 30,000 rpm, while the maximum engine
output speed required for an automobile is about 5,000 rpm. Hence, the
turbine shaft speed has to be geared down to a level practical for automobile
drive. This is accomplished by a fixed-ratio, reduction gear box installed
ahead of the transmission.
3-22
-------
1.6
1.4
% 1.2
a
.Q
^ 1.0
o
fe 0.8
CO
0.6
0.4
T» , TURBINE INLET
NONREGENERATIVE
CYCLE
TEMPERATURE = 1250 F
I
6 8
COMPRESSOR PRESSURE RATIO
10
12
Figure 3-10. Estimated Effect of Pressure Ratio and Turbine Inlet
Temperature on Fuel Economy for Nonregenerative
j 6 Cycle (Ref. 3-16)
Tt, TURBINE INLET
*l TCI JfMT B A -Tl ir»r- 1
1.4
ii.2
5
CD
0.6
0.4
ll TEMPERATURE
= 1250°F
REGENERATIVE
CYCLE
I
I
I
I
12
4 6 8 10
COMPRESSOR PRESSURE RATIO
Figure 3-11. Estimated Effect of Pressure Ratio and Turbine Inlet
Temperature on Fuel Economy for Regenerative
Cycle (Ref. 3-16)
3-23
-------
The function of the transmission is to convert the output from
the engine to useful levels of torque at the vehicle wheels. Gear-ratio
requirements for the gas turbine/transmission powertrain are dependent
upon the torque-speed characteristics of the particular engine type used. In
general, satisfactory vehicle operation demands high torque at low speeds to
provide rapid startup acceleration, while less torque is needed at the higher
vehicle speeds used for cruising. The transmission speed ratio range and
step size has to be selected so that these desired vehicle performance char-
acteristics are achieved.
The transmission requirements for different engine types may
be understood by referring to Figure 3-12, which shows the trend of torque
versus speed for the free (split-shaft) turbine engine and the fixed (single-
shaft) turbine engine. It may be seen that the torque curve for the free tur-
bine approximates an ideal characteristic, providing maximum torque at
low speed while maintaining a relatively high level of torque over the speed
range. The general characteristics of this curve are similar to the charac-
teristics of an internal combustion engine with torque converter. Hence,
conventional two-, three-, or four-speed transmissions may be used with
this engine.
The characteristic for the fixed turbine engine, on the other
hand, shows a very steep rise in torque over a narrow speed range. To
satisfy vehicle torque requirements with this engine demands the use of a
transmission with a wide range in gear ratios; a range of up to 10 to 1 in
eight or more steps may be required.
There are four major types of wide speed-ratio range trans-
missions under consideration for use with the single-shaft engine: the
hydromechanical, the 8-speed mechanical drive, the traction, and the belt
drive transmissions. The mechanical drive is a stepped variable speed
device; the other transmissions are stepless variable speed transmissions.
These other transmissions are considered to be more practical than the
b-speed mechanical drive for the single-shaft engine.
3-24
-------
200
% OF tnn
RATED 100
FREE TURBINE
TORQUE
FREE TURBINE
POWER
FIXED TURBINE
TORQUE
FIXED TURBINE
POWER
40 60 80 100
Figure 3-12. Typical Fixed and Free Turbine Torque and
Power Characteristics
3-25
-------
Considering the state of the art, the traction transmission
seems to offer the best available performance for use with the single-shaft
engine. One configuration of the traction transmission is capable of a total
of 9:1 speed ratio range (3:1 to 0. 33:1), which is adequate for single-shaft
engine operation. A startup clutch separates the transmission from the
gear box. The clutch engages when the transmission is put into the drive
position. The output of the transmission is connected to a standard torque
converter and gear box (Ref. 3-1). Figure 3-13 compares the fuel economy
of the single-shaft engine with a traction transmission and with a four-speed
automatic transmission. The advantage of the traction device is especially
evident at low speeds; for example, at 10 mph, the traction transmission
shows a 25 percent improvement in performance. However, the design
feasibility has to be proven in extensive tests before a single-shaft engine
can be selected for an automobile power plant.
3.2.3. Operating Characteristics
3. 2. 3. 1 Power Control Requirements and Methods
3.2.3.1.1 Control Devices
The turbine inlet temperature and the engine shaft speed have
to be controlled to prevent engine failure. In addition to these limit controls,
other control functions are required to ensure proper operation of the engine
at all times. For example, at part load the engine would operate at reduced
turbine inlet temperature, resulting in poor fuel economy unless certain
control devices, such as variable compressor-inlet-guide vanes and variable
turbine nozzles are used. These devices and the overall control system are
discussed in the following sections.
3- 2. 3. 1. 2 Compressor Variable Inlet Guide Vanes
The compressor performance is governed by three basic param-
eters: rotational speed, pressure ratio, and flow rate. At low flow rates,
compressor surge effects can produce a significant reduction in compressor
3-26
-------
16
14
12
TRACTION, TRANSMISSION
FOUR-SPEED
AUTOMATI
TRANSMISSION
NOTES:
1. ENGINE RATED POWER = 108 hp
2. 4-hp ACCESSORY LOAD
3. VEHICLE WEIGHT = 4000 Ib
10 20 30 40 50
VEHICLE SPEED, mph
60
70
Figure 3-13. Estimated Road-Load Fuel Economy of Single-Shaft
Regenerated Engine, 85°F Sea-Level Day (Ref. 3-1)
efficiency. Incorporation of variable inlet guide vanes alleviates these
problems to some degree. By adjusting the position of these guide vanes,
the air flow of the compressor can be reduced without affecting the rota-
tional speed and pressure ratio of the engine. The turbine-inlet tempera-
ture can then be maintained at or near the design point value, and this re-
sults in marked improvements in specific fuel consumption at part load.
3.2.3. 1. 3
Turbine Variable Nozzle
The use of a variable geometry nozzle in the power turbine
of the dual-shaft system permits changes to be made in the power split
between the power turbine and the first-stage turbine. This allows flexibility
3-27
-------
in control of the overall engine pressure ratio and operation at optimum
turbine inlet temperature. Similar to the compressor inlet guide vanes,
incorporation of the variable turbine nozzle improves the part load fuel
economy.
The braking effect obtainable from the variable nozzle feature
is illustrated in Figure 3-14, which shows braking force characteristics as
a function of car speed for two of Chrysler's free turbine, variable nozzle
engine designs (Ref. 3-17). The most recent Chrysler design, the sixth-
generation engine, was modified from the fourth-generation design to achieve
better braking characteristics. Also shown in the figure is the braking force
of a V-8 engine-powered vehicle with the transmission in the drive position.
As indicated, the braking power of the sixth-generation gas turbine engine is
substantially higher than that of the fourth-generation engine and slightly
better than the braking power of the standard automobile.
3. 2. 3. 1.4 System Controls
System controls involve the overall regulation of component
interactions during engine response to driver commands so that all critical
operating variables are maintained within safe limits. The major elements
of the control system could include the following items (Ref. 3-1):
• Fuel Control
• Turbine Inlet Guide Vane Actuators
• Compressor Inlet Guide Vane Actuators
• Automatic Starting and Limit Protection Control
• Transmission Control
• Hydraulic Clutch Control
• Power Boost Control
• Sensors.
Control operation could be hydromechanical or electronic. The major differ-
ence between the two systems is in sensors, metering sections, and controller
elements. In general, hydromechanical control systems use a rotating shaft
3-Z8
-------
500
bJJ
400
co
o
CO
o
300
uJ 200
o
OS
100
(50% Gas Generator Speed)
85° F Day
6th GENERATION ENGINE
WITH 1350°F POWER TURBINE
DISCHARGE TEMPERATURE
V-8 CLOSED
THROTTLE
TRANSMISSION
IN DRIVE
20
ROAD LOAD REQUIRED
B-BODY, 4250 Ib
4th GENERATION ENGINE
WITH 1200°F POWER TURBINE
DISCHARGE TEMPERATURE
30
40 50 60
VEHICLE SPEED, mph
70
80
Figure 3-14. Engine Braking Force Characteristics
(Ref. 3-17)
3-29
-------
and flyweights as the speed-sensing element and a controller to position a
metering valve. Electronic control systems have an electromagnetic speed-
sensing element, electronic controller elements, and electromechanical or
electrohydraulic metering sections.
A closed-loop control system functional schematic for a
variable inlet vane, single-shaft, gas turbine-powered vehicle is shown in
Figure 3-15. Driver command and operation would be identical to that for
existing automatic transmission vehicles powered by spark ignition engines.
In this design, the engine is coupled to the rear wheels through an infinitely
variable traction-type transmission and a forward/reverse-type gearbox.
Engine operation with zero vehicle speed is provided by a hydraulic clutch
that decouples the drive train from the engine. Essential elements in this
system include the engine fuel management components and transmission
control components. Although somewhat more complex, fuel management
of a gas turbine can be compared in function to a carburetor system in a
spark ignition engine. Other controlling elements include items such as
the inlet guide vane control and the power boost control (water injection) for
maximum acceleration.
The control system functional schematic for a free-turbine
powered vehicle is comparable to the single-shaft configuration; the major
difference is in the transmission control. This difference is a result of the
three-speed automatic transmission considered for this application and the
special requirements for preventing free-turbine overspeed.
Ford utilizes a double closed-loop control of engine speed and
turbine inlet temperature for automotive applications where both shaft power
and engine speed are varied. In this control system, predetermined sched-
ules of first-stage turbine speed versus power level and turbine inlet temper-
ature are built into the control system. The system senses first-stage
turbine speed and turbine inlet temperature and regulates the power turbine
nozzle angle to match the prescribed temperature. Optimum part-load fuel
3-30
-------
LIMITED SLIP
DIFFERENTIAL
PRESSURE
SIGNAL TO
ROLLER DRIVE
ACTUATION
SYSTEM
TURBINE ULTIMATE
'OVERSPEED SIGNAL
. TURBINE OVERTEMPERATURE
SIGNAL
Figure 3-15. Control System Functional Schematic,
Single-Shaft Gas Turbine Engine
(Ref. 3-1)
economy is maintained by operating at the highest turbine inlet temperature
consistent with the required compressor-surge margin. By adjusting the
nozzle vane position, the design-point, turbine-inlet temperature can be
maintained for engine speeds as low as 85 percent of the design value. Below
that point, the turbine inlet temperature is reduced to avoid compressor
surge (an unstable flow condition).
In case of overtemperature, the fuel supply to the engine is
reduced. Overspeed protection is provided by reducing the fuel flow and by
adjusting the nozzle vanes to the braking position.
3-31
-------
Chrysler's sixth-gene ration free turbine engine utilizes an
open-loop hydromechanical control system selected for cost reasons. Under
its current EPA contract, Chrysler will investigate closed-loop electronic
fuel control (EFC) system approaches. The principal advantage of closed-
loop EFC is its capacity to improve system response. In addition, the
turbine inlet temperature and the position of the power turbine nozzle can
be optimized in this system to provide minimum engine fuel consumption at
all operating conditions. Chrysler prefers EFC systems but is concerned
about the life of the temperature sensors (a problem not present in the
open-loop hydromechanical system). In its 50-car program, the sensors
failed after about 5, 000 miles as a result of oxidation of the thermocouple
beads and/or erosion of the shields.
3.2.3.2 Starting Requirements
Unlike a spark ignition engine, the gas turbine engine requires
a design for automatic control to assure engine protection during startup.
Engine protection is necessary because certain system malfunctions could
result in overtemperature or overspeed. These two conditions are not sig-
nificantly critical in a spark ignition engine, because combustion takes place
with the air fuel ratio less than stoichiometric. Thus, an increase in fuel
(failed carburetor) results in a decrease in combustor temperature. How-
ever, the opposite effect is true of the gas turbine engine. Also, overspeed
( a runaway condition ) is less likely in a spark ignition engine, since the
combustion chamber and valve train tend to limit the rotational speed of the
engine.
3. 2. 4 Packaging and Installation
3
Generally, the specific volume (ft /hp) of gas turbine engines
tends to be lower than that of spark ignition reciprocating engines. There-
fore, there should be no problem in fitting the engine package into a conven-
tional-sized engine compartment unless combustor or regenerator volumes
were to increase substantially. The engine could be located either in the
3-32
-------
front or in the rear of the vehicle. A front installation with front-wheel
drive or a rear installation with rear-wheel drive would provide improved
road traction and would eliminate the long driveline required in conventional
car designs. A rear installation would also eliminate the long exhaust ducts
and the resultant chassis redesign required to accommodate the duct volume.
Although the engine itself is compact, the air-intake and exhaust ducts are
large (the gas turbine has five times the air flow of an internal combustion
engine), and these have to be packaged properly to meet reasonable envelope
requirements and to provide sufficient ground clearance. Ford indicated that
packaging the turbine-exhaust system was a problem. Installation of gas
turbine engines in conventional vehicles required tearing up of the car-floor
pans. (Were the engine less sensitive to exhaust back-pressure, smaller
ducts could be used. ) Williams Research indicated that the exhaust system
for the turbine car could cost twice as much as for the conventional car.
3.3 PERFORMANCE CHARACTERISTICS
3. 3. 1 Exhaust Emissions
Both Chrysler and Williams Research have tested their gas
turbine automobiles for emissions over the Federal Driving Cycle (FDC).
Results of these tests are presented in Table 3-4. In these tests, the
engines were equipped with conventional state-of-the-art combustors. The
Chrysler results meet the CO and HC standards, but not the NO standard.
X
The two Williams Research gas turbine engines (WR-25 and 131-Q) installed
in a 1971 Hornet and a 1965 Volkswagen showed a high level of exhaust emis-
sions. These data were extracted from References 3-19 and 3-20. The
computed performance of other conventional combustors operated over a
simulated FDC are also listed in Table 3-4; none of these meet the 1976
NOX requirement.
A number of EPA contractors have been working on advanced-
design combustors. Based on experimental component testing, the emission
levels for engines equipped with these devices are predicted as shown in
3-33
-------
Table 3-4. Exhaust Emission Data over the Federal
Driving Cycle
Engine /Manufacturer
a. Conventional Combustors
United Aircraft
Research Laboratories
RGSS-6
United Aircraft
Research Laboratories
SSS 10
Chrysler sixth-
generation engine
Williams Research/
WR26/AMC Hornet
Williams Research/
131Q/Volkswagen
b. Advanced Combustors
Solar
General Motors
2 25 -horsepower
regenerative engine
AiResearch
United Aircraft of Canada
c. Clean Air Act Requirements
Emissions (gm/mi)
CO
0.53
1.86
3.99
7.43
6.92
4. 5
3. 3 4
2.4
1.6
3.64
3.4
HC
0. 15
0.31
-0.26
0.62
0.72
0. 34
0. 11
0. 015
0.67
0. 49
0.41
N0x
2.72
1.03
2.21
2.8
2. 5
1.81
0. 34
0. 315
0.44
0. 52
0.4
Remarks
Ref. 3-21 Calculated values
over simulated Federal Driving
Cycle, based on emission data
from GM engine GT-309.
Ref. 3-21 Calculated values
over simulated Federal Driving
Cycle, based on emission data
from GM engine T-56.
Ref. 3-31, 1975 Federal Test
Procedure emissions corrected
(background level subtracted
from measured values).
Cold start (Ref. 3- 1 9 test data,
Hot start < 1975 Federal Test
( Procedure
Ref. 3-20 test data, 1972 Fed-
eral Test Procedure.
Calculated values for simulated
Federal Driving Cycle, based on
advanced combustor data.
(Ref. 3-23)
Experimental test bed engine,
chassis dynamometer test with
5,000-lb car, 1975 Federal Test
Procedure (Ref. 3-18).
Calculated values for simulated
Federal Driving Cycle, for recu-
perated engine with 10 percent
bypass. (Ref. 3-25)
Calculated values over simulated
Federal Driving Cycle based on
advanced combuator data.
(Ref. 3-29)
1976 Federal standards
3-34
-------
Table 3-4. The exhaust emissions are generally close to or below the
original 1976 requirements. The AiResearch data, which simulates the use
of an advanced combustor in a recuperative engine with 10 percent recupera-
tor bypass, meets the CO and HC standards but is slightly above the NO
X.
requirement. These laboratory tests have proven the feasibility of reducing
exhaust emissions to meet the 1976 requirements.
A program sponsored by General Motors has resulted in
installing an experimental 225-horsepower gas turbine engine in a car meeting
the original 1976 Federal emission standards. This workhorse two-shaft
engine has single regenerator disc; it operates at a maximum gasifier
(compressor-turbine speed of 44,000 rpm, 4.5 atmospheres pressure ratio,
1850°F maximum turbine inlet temperature and specific fuel consumption of
0.55 Ib/bhp-hr. The test was conducted with an external control system that
has not yet been developed into an on-board system (Ref. 3-32).
3.3.2 Fuel Economy
The estimated SFC for design point operation of several
engines was tabulated in Table 3-1; design SFC values ranging from 0.4 to
0. 95 were shown. However, since automobiles operate over a range of
part-load conditions, the performance indicated at the design point is only
crudely representative of engine fuel economy over a normal driving cycle.
Fuel economy can be evaluated more realistically when engine
operation is simulated for an automobile run over the FDC. In the engine
optimization studies performed for EPA, the fuel economy of several gas
turbine engines was evaluated and compared with the fuel economy of a
representative V-8 spark-ignition engine.
Fuel economy values estimated by United Aircraft for single-
shaft, 150-hp gas turbine vehicles are shown and compared with a baseline
Otto cycle vehicle (OC-70) in Table 3-5. The major conclusion drawn from
this comparison was that the fuel economy of a regenerative or recuperative
engine was superior to that of a simple-cycle gas turbine engine or a spark
ignition engine.
3-35
-------
Table 3-5. United Aircraft Estimated Fuel Economy,
Baseline Vehicles (Ref. 3-21).
Ambient Temperature 59°F
Fuel Economy (mpg)
Engine
RGSS-6
(Regenerative
Single Shaft)
RCSS-8
(Recupe rative
Single Shaft)
SSS-10
(Simple Cycle)
OC-70 '
FDC
12.16
11.20
7.51
9.49
20 mph
15.06
13.73
8.86
10.65
30 mph
18.02
16.59
11.21
12. 62
40 mph
18.65
17.33
12.51
14. 10
50 mph
17.96
16.73
12.63
13.71
60 mph
16.59
15.42
12.22
12.91
70 mph
14.70
13.79
11.47
11.48
Vehicle Test*
Weight (Ib)
4000
3950
3700
4300
Simulated
'Olio Cycle
AiResearch predictions of fuel economy are shown in Table
3-6. Calculated results for two free-turbine, regenerated cycle, 135-hp and
175-hp engines, operating with three-speed automatic transmissions, are com-
pared with a spark ignition engine. Both FDC and fixed operating speed re-
sults are indicated. The fuel economy over the FDC is highest for the small
Table 3-6. AiResearch Estimated Fuel Economy of Free-Turbine
Regenerated Cycle (4000-lb Test Weight Vehicle,*
4-hp Accessory Load, Three-Speed Automatic
Transmission), (Ref. 3-1).
Level at
59'F Without
Boost, shp
175
135
175
Fuel Economy, mpg
Ambient
Temper -
ature,
°F
85
85
59
FDC
(1.3-hp
Accessory
Load)
16.4
18.1
12.5
20 mph
20.6
23.3
16. 1
30 mph
23.2
23.8
18.2
40 mph
23.3
25.9
18.2
50 mph
23.4
25.8
17.2
60 mph
22.3
24.3
15.5
70 mph
20.6
21.1
13.7
Simulated
3-36
-------
135-hp engine. This engine requires water injection to meet the maximum
power requirements. The fuel economy of the 175-hp regenerative engine,
while less than that of the 135-hp engine, shows a 31-percent improvement
over the spark ignition engine.
Existing gas turbine automobiles, however, show lower fuel
economy. The average fuel economy of the Chrysler sixth-generation gas
turbine engine installed in the int ;rmediate-size car is approximately 8.5
miles per gallon over the FDC. The peak fuel economy of the same car is
18 miles per gallon at 40 miles per hour (Ref. 3-17). The AMC Hornet with
the Williams Research WR-26 engine rated at 80 horsepower had a fuel econ-
omy performance range of 5. 1 to 8. 1 miles per gallon during several tests
conducted for exhaust evaluation using the FDC (Ref. 3-19).
3.3.3 Engine Noise and Vibration
The noise emanating from gas turbine engines is composed of
different frequencies with the spectrum varying with engine speed. The noise
level tends to be higher at idle speeds and lower at high speeds, when compared
to an internal combustion engine, because of the high-velocity flow of air at
the compressor inlet, high idle speeds, and interference effects due to the
variable inlet guide vanes. The gas turbine engine also lacks the damping
characteristics provided by the concentrated mass of the internal combustion
engine block. In general, the regenerative gas turbine engine may be expected
to have a lower noise level than the simple-cycle engine, because the heat ex-
changer in the regenerative case absorbs part of the sound energy in the exhaust.
A number of investigators expect that the automotive gas tur-
bine engine system will require exhaust-noise silencers (particularly for
simple-cycle engines). The present Chrysler engine package is reported
not to require exhaust mufflers.
Chrysler expects that a number of design modifications would
be required in the gas turbine vehicle's chassis and frame. These include a
new suspension system, a new exhaust system arrangement, and a new front
end. The new suspension would be of the rubber-insulated type to provide a
quieter ride.
3-37
-------
According to drivers in the Chrysler 50-car test program
(Ref. 3-22), the gas turbine engine operated with significantly less vibration
than the conventional piston engine. This is directly attributable to the pure
rotary motion of the engine.
3.3.4 Maintenance and Durability
Modern piston engines operated in normal passenger car
service as expected to accumulate 75,000 to 100,000 miles without requir-
ing a major overhaul. The impact of exhaust emission controls on this type
of durability (and the expected reliability) is still under evaluation. The gas
turbine engine, with fewer parts may, assuming proper care and mainte-
nance, provide even higher durability.
Contractors involved in the EPA-sponsored engine configu-
ration optimization study addressed the problem of gas turbine car mainte-
nance and concluded that scheduled maintenance and service requirements
(chassis lubrication, air and oil filter changes, etc.) would be much the
same as for the conventional automobile. AiResearch, for example, esti-
mated that engine-related repairs, although different in nature, would occur
at about the same frequency among the aggregate of service or repair items
for both engine types.
In the Chrysler 50-car program (Ref. 3-22), 1. 1 million
vehicle miles were accumulated in a period slightly over 2 years. In the
beginning of this program, the lost time due to engine malfunction was about
4 percent; this was eventually reduced to 1 percent as familiarity with re-
pair problems was acquired. The Chrysler experience suggests that the
training of mechanics in maintenance and repair of gas turbines will not
present unusual problems.
Chrysler anticipates that production models of gas turbine
engines will require less maintenance than their internal combustion engine
counterparts. This judgment is partly based upon experience with the sixth-
generation prototype engine, which was serviced after 3, 500 hours of oper-
ating time, corresponding to a mileage accumulation of about 170,000 miles.
A potentially serious service problem in regenerated gas tur-
bines concerns the high frequency of maintenance encountered with the
3-38
-------
regenerator seal, a seal problem that must be solved before this engine type
can be placed into mass production.
3.3.5 Safety
Aircraft and stationary gas turbine engines have an excellent
record of safety performance, due largely to the practice of employing reg-
ularly scheduled preventive maintenance and overhaul. In the passenger car
application, systematic maintenance cannot be guaranteed. Accordingly,
special safety provisions must be incorporated into the system's basic design
to prevent severe damage in the event of catastrophic failure, as follows:
• Containment has to be provided so that in case of engine
runaway or turbine blade failure, the rotating elements of
the engine will be safely confined.
• Limit controls must be designed so that engine operation
will not exceed allowable limits.
• Backup provisions, such as engine packaging and installa-
tion arrangements, must be made so that the vehicle occu-
pants are not subjected to heat or other hazards in the
event that primary safety provisions fail under high-
impact accident or other catastrophic conditions.
No fundamental difficulties are anticipated in meeting these objectives for
the passenger car turbine.
3.3.6 Driveability
In general, the gas turbine car does not have the standing-
start response of a piston-powered vehicle, circa 1970. This is exemplified
by the performance characteristics shown in Figure 3-16 where the standing-
start acceleration performance for a regenerated free turbine engine vehicle
and a single-shaft turbine engine vehicle are compared with a spark ignition
engine vehicle. As shown, it is estimated that the single-shaft system has
up to a 1-second delay in starting vehicle motion, and the free turbine has up
to a 0.4-second starting delay; the precontrolled emission, spark ignition
engine has a delay of less than 0. 1 second. *
Some comments on driveability by drivers participating in
Chrysler's 50-car test program, as noted in Reference 3-22, are as follows:
(1) The operation of the engine was smooth and vibrationless. Passengers
3-39
-------
UJ
200
180
160
140
120
100
80
60
40
20
0
FREE TURBINE
REGENERATED,
THREE-SPEED
AUTOMATIC,
—175 hp RATED
(est.)
TIME, sec
NOTES:
1. TAMB = 105° F
2 VEHICLE WPIRHT = ^OOO-'b TURBINE ENGINE
d. VEHICLE WEIGHT = 4300.,b 5, ENG|NE
Figure 3-16. Standing-Start Acceleration
Performance (Ref. 3-1)
felt a gliding sensation at all speeds, which was particularly pleasant on long
trips. (2) Starting ability was said to be superior to that of the internal
combustion engine. Users consistently stated that the turbine car was supe-
rior to conventional cars in providing fast and sure ignition. (3) Other supe-
rior operating features mentioned were good engine power, low vibration,
and nonstalling characteristics. (4) Users tended to contradict each other
in reaction to the sound of the turbine engine. For every person who com-
plained about the noise level, there were three or four who liked the sound.
(5) The main negative comment was lack of acceleration, primarily when
starting from standstill. Chrysler attributed this to the relatively low
engine design power level in this automobile (130-hp).
3-40
-------
Chrysler improved the acceleration and braking response
characteristics of their sixth-generation engine, relative to the fourth-
generation engine performance, by shifting major accessories from the
first-stage turbine shaft to the power turbine shaft. Braking response
characteristics were improved with an increase in turbine inlet tempera-
ture, providing additional braking power at negative turbine nozzle blade
incidence. These improved characteristics were displayed in Figure 3-14.
3.3.7 Starting Characteristics
The time required from driver initiated engine startup until
power is available for vehicle acceleration is a performance factor. The
cold-engine, start-time characteristics of the Chrysler sixth-generation auto-
mobile gas turbine (see Figure 3-17) show that starting time is significantly
affected by the ambient temperature.
3.4 CURRENT STATUS OF TECHNOLOGY
3. 4. 1 Current Utilization
As pointed out before there are no gas turbine automobiles in
production today. The only engine considered as a possible candidate for
production was the Chrysler engine in their well publicized 50-car program.
The engine specification and design point characteristics are shown in
Table 3-7 (Ref. 3-22).
During the Chrysler 50-car, gas-turbine test program, which
lasted for 2 years (October 1963 to January 1966), cars were assigned to
selected individuals and, as reported by Chrysler- the reaction was gener-
ally favorable. However, in 1966 Chrysler decided to abandon their origi-
nal idea of a low-volume-production, gas turbine vehicle because of a num-
ber of deficiencies uncovered during the program. These included poor fuel
economy, inadequate engine response and braking, insufficient engine power,
and excessive noise. They have continued gas turbine work as a research
program, as described in the next section.
3-41
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0'
-20
TEST TIME TO 65% POWER
TEST TIME TO IDLE SPEED
20 40 60
AMBIENT TEMPERATURE, °F
80
Figure 3-17. Starting Time versus Ambient
Temperature (Ref. 3-17)
3-42
-------
Table 3-7. Specifications of Chrysler Corporation's Fourth-Generation
Gas Turbine Engine (Ref. 3-22)
General
Type:
Regenerative gas turbine
Rated Output:
Power - 130 bhp @ 3, 600-rpm output shaft speed
Torque - 425 Ib-ft @ zero-rpm output shaft speed
Weight:
410 Ib
Basic Engine Dimensions (without accessories):
Length - 25 inches
Width - 25.5 inches
Height - 27.5 inches
With automotive accessories in place, the overall length is:
35 inches
Fuels:
Unleaded gasoline, dies el fuel, kerosene, JP-4, etc.
Components
Compressor:
Type - Centrifugal
Stages - One
Pressure Ratio - 4:1
Efficiency - 80%
3-43
-------
Table 3-7 (continued)
First Stage Turbine:
Type - Axial
Stages - One
Efficiency - 87%
Second Stage Turbine:
Type - Axial
Stages - One
Efficiency - 84%
Regenerator:
Type - Two rotating disks
Effectiveness - 90%-f
Burner:
Type - single can, reverse flow
Efficiency - 95%
* DESIGN POINT CHARACTERISTICS
Maximum Gas Generator Speed - 44, 600 rpm
Maximum Second Stage Turbine Speed - 45, 700 rpm
Maximum Output Speed (after reduction gears) - 4, 680 rpm
Maximum Regenerator Speed - 22 rpm
Compressor Air Flow - 2.2 Ib/sec
First Stage Turbine Inlet Temperature - 1700°F
Exhaust Temperature (full power) - 525°F
* Ambient conditions:
Temperature - 85°F
Barometric Pressure - 29.92 in. Hg
3-44
-------
3. 4. 2 Current Research and Development
3. 4. 2. 1 Engine Research and Development
The main EPA-sponsored engine research program is -with
the Chrysler Corporation.
Research and development on the automotive gas turbine has
been under way at Chrysler since 1950. The current engine is a modifica-
tion of their fourth-generation engine used in their 50-car test program.
The engine is a free-turbine regenerative design consisting of the following
basic components: (1) a single-stage radial compressor (4. 1:1 design pres-
sure ratio, 44,600-rpm design speed) driven by a single-stage axial turbine;
(2) two rotating regenerators; and (3) a separate variable nozzle, single-
stage axial power turbine (45,700-rpm design speed). The engine is con-
nected to the vehicle transmission through a reduction gear arrangement.
Axial turbine wheels were selected by Chrysler for a number of reasons,
including improved packaging and low inertia.
The principal new features of the current engine are a higher
first-stage turbine inlet temperature-- 1, 850° F vs 1,750°F during maximum
steady-state operation; 2, 000°F vs 1, 850°F during maximum acceleration,
a higher power output (150-hp vs 130 hp)--and a redesigned accessory drive
arrangement. On this current engine, major accessories (power steering
and alternator) are driven by the power turbine. In the fourth-generation
engine, all accessories were driven by the gas generator. The shift of major
accessories to the power turbine resulted in a faster response for the engine,
as shown in Figure 3-17- This change reduced the first-stage turbine load,
permitting a reduction in the size and number of high-speed bearings
required.
This 150-hp, sixth-generation engine, shown schematically in
Figure 3-18 and pictorially in Figure 3-19, is a research and not a production
engine. The fuel economy and the power output of the engine were increased
from the fourth-generation values by increasing the first-stage turbine inlet
temperature by 100°F. This was accomplished by fabricating the turbine
3-45
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AIR CONDmONINO
COMPRESSOR
AIR
INTAKI
AIR COMPRESSOR
BURNER
REGENERATOR H
IHIAT IXCHANSII) (J
REOINIRATOR
TO REAR WHEELS
Figure 3-18. Schematic of Chrysler Sixth-
Generation, Gas Turbine
Engine (Ref. 3-17)
3-46
-------
Figure 3-19.
Chrysler 150-hp Gas Turbine
Engine (Ref. 3-17)
-------
wheels from INCO 713C (70-percent nickel). Additional illustrations of
Chrysler's gas turbine engine installed in automobiles are given in Fig-
ures 3-20 through 3-23.
In December 1972, Chrysler was awarded a $6. 5 million
contract by EPA for the development of an improved version of the sixth-
generation gas turbine engine with respect to emissions and fuel economy.
Tests conducted to date with the 1972 Federal Test Procedure indicate that
the original 1975 Federal emission standards can be met by the sixth-
generation engine. However, NO emissions of 2.7 gm/mile obtained on
this engine are considerably above the requirement of 0.4 gm/mile. Reduc-
tion of NO is sought primarily by means of combustor modifications. Both
in-house combustor designs and combustors designed under separate EPA
contracts will be tested by Chrysler. The NASA Lewis Research Center is
playing an important role in the effort in the areas of combustion, rotating
machinery design, and engine test.
In addition to the U.S. research and development described
above, a number of foreign activities in automotive gas turbine development
are in progress. Volkswagen, for example, has undertaken a long-range
program to develop a passenger car gas turbine with performance and pro-
duction costs comparable to the spark ignition engine but with improved ex-
haust emission characteristics. This program is in the preprototype hard-
ware development stage. A research engine developed with Williams Re-
search Corporation, the VW-GT 70, has been built and tested. This is a
two-shaft, 75-hp machine with preheating and an integral reduction gear
designed to fit into the rear engine compartment of various Volkswagen
vehicles. Emissions over the Federal Driving Cycle for the VW-GT 70
are reported to meet the original U.S. standards for 1975; however, NO
3C
emissions are still significantly higher than the 1976 0.4 gm/mi level.
3.4.2.2 Components Research
3.4.2.2.1 Combustors
Development work conducted on gas turbine combustors has
indicated that sufficiently low HC and CO emission levels can be achieved
3-48
-------
TOP VIEW
SIDE VIEW
Figure 3-20. Chrysler Baseline Turbine-Powered
Vehicle (Ref. 3-30)
-------
I
I
3
Figure 3-21.
Chrysler 1 50-hp Gas Turbine
Engine, Vehicle Installation
(Ref. 3-30)
-------
STARTER
uo
Ui
SINGLE REGENERATOR
Figure 3-22. Chrysler 100-hp Upgraded Gas Turbine Engine with Single Vertical
Regenerator Compact Vehicle Installation (Ref. 3-30)
-------
COMBUSTOR
OJ
I
rv
ACCESSORY CLUSTER
DUAL AIR
INLETS
ACCESSORY CLUTCH
Figure 3-23. Chrysler 100-hp Upgraded Gas-Turbine Engine with Single-Shaft
Vertical Regenerator Compact Vehicle Installation (Ref. 3-30)
-------
with conventional combustors to meet the federal emission control objectives
for light duty vehicles. However, the levels of NO emitted by these com-
5t
bustors are considerably higher than the original 1976 0.4 gm/mi require-
ment. In conventional combustors, the fuel is injected at high pressure
through fuel injectors, and fuel atomization is achieved in the dome section
of the cylindrical or annular combustion chambers by means of jet impinge-
ment. Combustion of the fuel takes place in the near-stoichiometric fuel
vapor/air region surrounding the individual fuel droplets. As a result, the
combustion temperatures are very high, especially in the case of regener-
ated or recuperated gas turbine engines. Since the kinetically controlled
NO formation rate increases exponentially with temperature, the NO emis-
X X
sions from a conventional droplet-burning-type combustor are also very high.
For instance, Chrysler's sixth-generation gas turbine installed in a medium-
size vehicle has NO emissions of about 2. 2 gm/mi over the Federal Driving
Cycle. The NO emissions of nonregenerated gas turbines with conventional
.X
combustors tend to be somewhat lower, because of the lower temperature of
the incoming air.
In order to obtain satisfactory fuel economy in gas turbines,
some degree of regeneration appears to be required. Since conventional gas
turbine combustors are inherently high NO emitters it appears that new
3£
combustors would have to be developed if the gas turbine automobile is to
meet federal NO objectives. In view of current industry efforts in the
development of high-temperature gas turbines utilizing ceramic components,
the need for low NO combustors becomes even more important.
The concept of lean combustion of fully vaporized and pre-
mixed air fuel mixtures offers a potential solution to the NO emission prob-
3C
lem with current gas turbine engines. New combustor designs based on
this particular concept are now in the development stages. Although con-
siderable progress has been made to date by industry, much more remains
to be done, particularly in the area of durability and off-design operation.
A number of these advanced configurations are briefly described in the fol-
lowing paragraphs.
3-53
-------
•A*
SolarJ under contract to the EPA, has been involved in the
development of a low NOX, jet-induced circulation (JIG) combustor concept
for potential use in automotive gas turbines (Figure 3-24). In this design, the
fuel and the primary zone air are premixed before entering the primary
combustion zone through a variable nozzle-like opening. Emission levels
obtained during testing over a simulated Federal Emission Test Driving
Cycle indicate the feasibility of meeting one-half of the original 1976 Federal
Emissions Standards. Further development of the combustor and nozzle
activator mechanism is needed before a combustor/engine combination could
be used to power a vehicle.
In the Ford external vaporizing combustion (EVC) concept,
the fuel enters the premix/vaporization chamber through a coaxial nozzle
and is then atomized and premixed by the high-velocity primary air flow
(Refs. 3-4 and 3-5). Based on initial test results, Ford feels that the emis-
sion control objectives for light-duty vehicles can be met with a gas turbine
utilizing the lean premixed/prevaporized EVC concept. However, a number
of potential problem areas must be solved before the combustor is ready for
use. These include combustor stability and response under transient operat-
ing conditions, preignition, durability, and cold-start emission characteristics.
Aerojet has designed a novel combustor for potential use in
the EPA automotive gas turbine program. In this concept (Ref. 3-24) the
conventional fuel-spray nozzles are replaced by the premixer hardware
which distributes the initially liquid fuel, then injects it into the combustor
air through a large number of very small orifices (Figure 3-25). Based on
analytical work on flow mixing and combustion kinetics, Aerojet feels that
emission control requirements can be met with the platelet combustor.
AiResearch, under contract to EPA, has been involved in the
development of low-NC>x combustors utilizing recuperator bypass (Ref. 3-25).
A portion of the recuperator bypass air is injected through the dome section
Solar Division of International Harvester Company
3-54
-------
LAYOUT OF PHASE IIIIC-B COMBUSTOR
ih SOLA*
IIC-B VARUBU GEOMETRY LOW EMISSION COMBUSTOR
*SOUl»
IIC-B YSRIJBLE GEOMETRY LOW EMISSION COMBUSTOR
of
Figure 3-24. Solar Jet-Induced Circulation Combustor,
JIC-B(Ref. 3-23)
3-55
-------
FUEL
AIR FOR COMPRESSOR
AND/OR REGENERATOR
REACTION ZONE
RECIRCULATION ZONE
MIXING AND EVAPORATION ZONE
^ RECTANGULAR NOZZLES
BLUFF BODY
FLAME HOLDER
Figure 3-25.
Schematic of Aerojet Platelet
Premixer Concept (Ref. 3-24)
3-56
-------
of the combustor, the remainder of the bypass air and fuel are injected in
an upstream direction through an L-shaped pipe located downstream of the
dome (Figure 3-26). Although the emissions of this vaporizer combustor
with bypass are rather low, further reduction of NO is required before the
J\.
federal standards can be met in a gas turbine automobile.
In another EPA contract effort, General Electric worked on
the development of a low-emission porous plate combustor for use in auto-
motive gas turbine engines. In this concept (Ref. 3-26), the fuel is preva-
porized and the air fuel mixture is then discharged through the porous com-
bustor plate and ignited at the front face of the plate (Figure 3-27). Emission
measurements with propane/air mixtures were reported to be favorable
(Ref. 3-26), with NO and HC generally below the original 1976 Federal
.X
emission standards. CO was below the standards at combustor conditions
corresponding to light engine loads but above the standards at heavier loads.
In catalytic combustors a catalyst material, deposited on a
suitable ceramic substrate, is used to promote and sustain the oxidation
reactions at very lean air fuel ratios. Preliminary test data indicate
catalytic combustors are inherently low NO , CO, and HC emitters. How-
5C
ever many potential problems remain to be solved. These include catalyst
and substrate life, catalyst poisoning characteristics, temperature capa-
bility, transient response, maximum heat-release rate, and air fuel mix-
ture preparation.
3.4.2.2.2 Ceramic Components
Fuel consumption of gas turbine engines decreases as tur-
bine inlet temperature is increased. The specific fuel consumption of a
highly regenerated engine operating at a compressor-pressure ratio of from
4:1 to 6:1 can be improved by about 20 percent by increasing the turbine in-
let temperature from 1, 800° F to 2, 500° F. Based on current materials tech-
nology, turbine inlet temperatures are limited to about 1,800°F to 1,900°F
using nickel alloys. Higher temperatures can be achieved with transpiration-
cooled, veil-cooled, or internally cooled turbine blades. However, these
3-57
-------
L-PIPE
BYPASS AIR
AND FUEL
OJ
I
00
DOME
BYPASS AIR
DOME
PRIMARY ZONE
DILUTION ZONE
Figure 3-Z6. AiResearch Recuperator Bypass
Combustor (Ref. 3-25)
-------
Oo
I
(Ji
DILUENT AIR IN -—
FLAME FRONT
FUEL/AIR IN
COOLINGAIRIN
DILUENT AIR IN
i * > » i -f
> »^ > » »
FLAME FRONT
COMBUSTION
GASES OUT
COMBUSTION
GASES OUT
Figure 3-27 . General Electric Low-NOx
Porous-Plate Gas Turbine
Combustor (Ref. 3-26)
-------
approaches are limited to turbine inlet temperatures of about 2,300°F and
are not currently considered feasible for automotive gas turbines because
of high manufacturing costs. For these reasons, gas turbine manufacturers
have been interested in ceramic turbine nozzles and rotors for some time.
Although initial acceptance of these brittle materials has not been good, there
are indications that the properties of ceramics have now been improved to a
point where a number of manufacturers are seriously interested in the use
of these materials in their gas turbines.
No serious difficulties are foreseen in the use of ceramics
for stationary components of the engine such as combustor liners, nozzles,
shrouds, heat exchangers, and insulation. However, the design of rotating
components such as ceramic turbine wheels represents a very difficult
challenge. The wheel must be capable of withstanding the high thermal and
centrifugal loads and must be sufficiently light to provide adequate engine
response.
In June 1971, Ford, with Westinghouse as subcontractor, was
awarded a $10. 3 million ARPA-sponsored contract to demonstrate that
brittle materials can be successfully used in high-temperature gas turbine
components such as turbine inlet nose cones, nozzle vanes, turbine wheels,
and combustor liners. Under terms of the contract (which is monitored for
ARPA by the Army Materials and Mechanics Research Center) Ford is con-
centrating on development of ceramics for use in a small automotive gas
turbine vehicle, and Westinghouse is investigating the applicability of ceram-
ics to large stationary gas turbines (Ref. 3-27).
To date, Ford has concentrated its efforts in the areas of
materials technology, turbine component development, and stress analysis.
In the material technology area, Ford initiated a program to determine the
physical properties of various ceramics, including hot-pressed silicon car-
bide, hot-pressed silicone nitride (Si3N4), and injection-molded silicon ni-
tride. Based on these evaluations, Ford feels that hot-pressed silicon ni-
tride and silicon carbide are candidate materials for turbine rotors.
3-60
-------
Although very encouraging results have been achieved to date
in the ceramics area, considerably more development work is required before
these ceramics can be considered feasible for use in the manufacture of rotat-
ing components for applications to the automotive gas turbine engine. More
data are needed on material behavior as a function of time, temperature, and
method of fabrication. The effect of fuel impurities (e. g. , sulfur) should
be evaluated. Standardization of property testing and reporting is needed.
Product uniformity must be developed. The designer will have to learn to
cope with the behavior and limitations of ceramics, including lack of ductility
and sensitivity to stress concentrations. Attachment to metal components
with different thermal expansion properties requires further development.
Finally, the ceramic components must demonstrate the ability to withstand
many start/stop, acceleration and deceleration cycles characteristic of
automotive operation.
3.4.2.2.3 Water Injection
By injecting water or other suitable fluids into the engine air-
flow, the thermodynamic and physical properties of the working fluid are
altered so that more -work is derived per pound of flow. Injection could
occur at the compressor or combustor inlets. Analysis has shown that water
injection (water-to-air mass ratio of 0. 06) at an ambient temperature of
about 80°F, can increase engine power by 35 percent. However, this power
increase is obtained at the cost of a reduction in compressor efficiency. At
a water-to-air mass ratio of 0. 06 the ratio of compressor efficiency with
water injection to the compressor efficiency without water injection is 0. 83,
a substantial drop in compressor efficiency (Ref. 3-28). A positive result
of the power-boost feature is that the engine size can be reduced, resulting
in an improvement in part-load operating efficiency. NASA is supporting
Chrysler's efforts to evaluate this concept under the EPA contract.
3-61
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3. 5 PROJECTED STATUS
3. 5. 1 Potential for National or Regional Transportation
EPA studies conducted by engine manufacturers indicate that
the gas turbine has good potential for use as a future automobile engine.
However, automobile manufacturers have to be convinced of the economic
advantages of the gas turbine over the spark ignition engine before it can
become a strong candidate for future automobiles. In addition, the follow-
ing technical problems have to be solved.
• The efficiency of the engine has to be increased to improve
fuel economy. Development of a ceramic turbine will aid
in solving this problem by allowing a significant increase
in the turbine-inlet temperature.
• Emissions have to be lowered below the original 1976 Fed-
eral emission standards. Continued development of ad-
vanced combustors is the major approach taken to solve
this problem.
• Additional improvement and development of ceramic mate-
rials is required to enhance their suitability for regener-
ator, turbine, and other high-temperature, high-stress
applications in the gas turbine engine.
• Techniques have to be developed for mass production of
cast, high-temperature, super-alloy components.
• Methods have to be developed to verify low production
costs for control systems and regenerators.
It is anticipated that a period of from 7 to 11 years will be required before
gas turbines can be mass produced for automobiles. A follow-on to the
current EPA program with Chrysler can be conducted to develop a low-
production version of an automobile gas turbine engine. As additional re-
search and development results become available, this engine can be modi-
fied if production processes permit.
3.5.2 Impact on Energy Requirements and Air Quality
The Chrysler baseline engine program is directed toward
proving that significant reductions in exhaust emissions and major
3-62
-------
improvements in fuel economy are practical with gas turbine engines.
This new engine could then be a major contributor to the reduction in air
pollution.
With the development of high-temperature turbine materials
and operation at 2,500°F, the efficiency of the gas turbine cycle should in-
crease to the point -where a 20 percent fuel economy improvement over cur-
rent gas turbine road performance can be anticipated. This would allow gas
turbine-powered automobiles to be better than currently projected for piston-
engine cars controlled to the original 1976 Federal emission standards.
High-octane fuels are not necessary for the gas turbine; thus,
the issue of fuel lead additives as a potential air pollutant and toxilogical
hazard would be eliminated as a concern with the use of these engines. In-
deed, vehicular gas turbines can use a variety of fuels, ranging from natural
gas to diesel and heavier oils, thereby providing the potential for making
maximum use of available fuel resources.
3-63
-------
SECTION 4
-------
4. RANKINE CYCLE ENGINES
4.1. INTRODUCTION
4. 1. 1 General Description
The Rankine cycle engine is an external combustion engine in
which high pressure steam or some other working fluid is expanded in a recip-
rocating (piston-type) or turbine-type device to produce work. Figure 4-1
shows the components of a typical Rankine cycle engine in schematic form.
The feed pump draws liquid at a low pressure from the reservoir and forces
it under high pressure into the vapor generator where it is converted to super-
heated vapor by heat from the combustion gases. The hot, high pressure vapor
then flows into the expander where it gives up energy as it expands to a lower
pressure and temperature. Exhaust vapor from the expander enters the con-
denser where it is converted back to a liquid by rejecting heat to the surround-
ings. The resulting low pressure liquid is then returned to the reservoir or
sump to complete the cycle. The major components shown (feed pump, vapor
generator, expander, condenser) are required for any Rankine cycle system.
Specific systems will show variations in such features as type of drive for
the auxiliaries, additional heat transfer paths from one part of the cycle to
another, and the type and location of liquid reservoir and boost pump.
4.1.2 Historical Background
Rankine cycle steam cars held a prominent place in the early days
of the automobile. Steam cars were sold commercially in the late 1800s and
early 1900s with familiar names such as Stanley Steamer; Locomobile, White
and Doble. These early steam cars had a simple pressurized boiler containing
a large quantity of water which required about a 30 minute cold startup, and
posed a continual safety problem. They were noncondensing, dumping exhaust
steam overboard. Consequently, frequent stops were required for makeup
water. In spite of these disadvantages some of these cars, such as the
4-1
-------
ro
ATOMIZING
AIR
PUMP
HYDRAULIC
PUMP
ELECTRIC
MOTOR
VAPOR
GENERATOR
iXl
THROTTLE
ACCUMULATOR
HYDRAULIC
AIR PUMPS
INTAKE
AUXILIARY!
POWER
HYDRAULIC
MOTOR
EXHAUSTJI
HYDRAULIC
MOTOR
J I
I
FEED
PUMP
TRANSMISSION,
DRIVE WHEELS
FAN
HYDRAULIC
MOTOR
AIR-
COOLED
CONDENSER
SUMP
SUBCOOLER
MAKEUP
PUMP
Figure 4-1. Typical Rankine Cycle System, Schematic
-------
Stanley Steamer, were quiet-running cars of quite satisfactory performance.
Various technical advances were made with the years, primarily the monotube
boiler by Doble, which effectively solved the safety and startup problem.
Condensing radiators were in use from about 1915 on, resulting in true closed
Rankine cycle operation. By this time, however, the internal combustion
engine had entered its era of low-cost mass production. In this respect the
steam car did not compete, disappearing from the automotive scene in the
very early thirties.
Interest in the steam car occurred sporadically during the
fifties and early sixties, but it was not until the latter sixties, with the impact
of impending air pollution control legislation becoming apparent, that the
Rankine cycle automobile system began to receive serious scientific attention.
The impetus behind this renewed interest lay in the clean burning characteristics
of the external combustor of this system, which requires neither post-engine
exhaust cleanup devices, nor compromises with an efficient burning process.
A number of engineering efforts were undertaken at about this
time by widely diverse organizations. This development received impetus and
cohesion with the initiation in 1970 of the EPA Advanced Automotive Power
Systems (AAPS) Program. The purpose of this program was, first, to identify
the key problem areas and development tasks, and then to proceed with experi-
mental and analytical programs which would result in an accurate appraisal of
the overall feasibility of the Rankine cycle engine as a practical, mass pro-
duced power plant. Table 4-1 is a summary of the more significant world-
wide Rankine cycle efforts which are either completed or now under way.
4. 2 POWER PLANT DESCRIPTION
A Rankine cycle engine is one in which the heat generated by
combustion is transferred to a second fluid, which then does work on mechan-
ical components to provide engine power. A major reason for current interest
in Rankine cycle engines is based on this external (to the working fluid) com-
bustion. Since the combustor is required to provide heat only, combustion
pressures can be atmospheric; maximum combustion temperatures can be
4-3
-------
Table 4-1. Summary of Rankine Cycle Engine/Vehicle
Development Programs
Ur\ploprr/ronlra< tor
M'A AAPS Program
A«-r.iir( Liquid I< • u krl Co.
] n,,r M..IMTS r.,,rp.
Scientific I ncrcy SyslcmR
1 hr-rm.j Hei 1 ron Corp.
hrobn k *. Associates
Lear Motors < or p.
Sieam Power Synterns Co.
UOr/Dallas Program
Sundstrand Co rp.
California Clean Car Project
Aero let Liquid Rocket Co.
Steam Power Systems Co.
Other Domestic Programs
Kinetics, Inc.
Williams Engine Co.
Foreign Programs
Prilchard Steam Power Proprietary, Ltd.
SAAB-SCANLA
Politecnico Mila.no
'Simulated or installation design target.
V rh i c 1 e
li-Passcnger car3
S- Passenger car3
51 -Passenger bus
25 - Pas se-nge r bus
4-Passenper car3
4 -Passenger car
(\'olkswagen|
4-Passenger car
4-Passenger car
6- Passenger car
Busa
bh - b h
p
Rated (or Estimated)
150,K\pander ^ross
1 23, Kxpandor gross
158. F.xpandr r ^ross
1 38. T ransmission in
Mb, Kxpandcr gross
240, Cross bhp
200. Net bhp
240. Cross bhp
180. Net bhp
275, Gross bhp
224, Net bhp
90, Gca r box mil
faO. Expander gross
65. Expander gross
70
160
100
sepower DOT =
Type of Expander
and Rated Speed
in rpm
Turbine; 32,000
Turbine; h^.OOO
Piston; :. 500
Piston; 1 , 800
Piston; 2. 1 00
I urbinr; (i5. 000
Piston. 1, 850
Turbine; 60,000
Piston; 2,400
Piston; 5, 000
Gerotor type
Piston
Piston
Piston
Turbine
W.irkinc Fluid
AEF-7H (Proprietary)
U -urr
V\ a 1 r r
1- lunrmol - 85
\\' a I e r
Water
Water
CP-25; a single ihemical
compound comprised of
methyl ben/cne, loKicnt-,
and toluall.
Water
\\' a t e r
Water
Refrigerant I 13
Water
Water
Water
Organic
P P
Status of Program
Preprototypc engine devcl-
npmrnt program ^ ompletcd
in December pi? ^. One
rontrai-tor (5KS) sele. trd to
pmcrcH with pr.>tntypr rn«-
inr development . In) lower!
by insl A 1 lati«n in \ehulr.
ember l'»72. (onsisied
of demnnst r at ion use in
puhlu service and vehicle
emissiong testing.
Program completed.
Consisted of service in
transit system and vehicle
emissions tesling.
Testing and vehicle instal-
lation in progress, largel
date for vehicle delivery
is May 15, 1974.
Vehicle first driven March
ments being made.
No published emissions or
fuel economy data.
Recently initiated.
-------
relatively low without strongly affecting cycle efficiency, and the flow rate of
the combustion gases can be relatively steady (noncyclic). Under these inherent
conditions, there is good evidence that efficient engine systems can be developed
which emit very low air pollution.
4.2.1 General Power Plant Configuration
The basic ideal Rankine thermodynamic cycle for water is shown
schematically in Figure 4-2. Starting at point A, the subcooled liquid is
pumped up to boiler pressure, point B, with a feed pump. Heat from the
combustor is then added at constant pressure in the boiler. The three main
thermal portions of the boiler are shown as the heating section (line B - C),
boiling (line C - D), and superheating (line D E). The steam leaves the
boiler at pressure and temperature conditions, E, and enters the expander.
Work is extracted in the expander by rotary or reciprocating mechanical
devices, and the steam conditions drop to those at point F which for purposes
of demonstration is shown as still lying in the superheat region. In some
systems, particularly those using organic working fluids (see Section 4.2.2.6),
some of this remaining heat is transferred to the working fluid (from process
line F - G to process line B - C) in a regenerator. The water then enters
a condenser, at point G, where sufficient heat is removed and rejected to
atmosphere to condense all of the remaining steam at constant temperature
(line G - H) and to accomplish some subcooling of the liquid (line H - A).
Very early steam engines often simply dumped the steam from the expander
(at point F) to atomsphere in a so-called open cycle and drew fresh water
from a stored supply at condition A (letting the environment close the cycle).
All of the modern practical systems for automotive vehicles close the cycle
through the condenser, as shown in Figure 4-2, to conserve the working fluid
and thereby maintain acceptable vehicle travel range.
Basically, the work derived from the cycle of Figure 4-2 is
the difference between the areas under the curves A - E and F - A. Efforts
4-5
-------
• E
t
QC
UJ
O_
ENTROPY
Figure 4-2.
Typical Ideal Rankine Thermodynamic Cycle with
Water as a Working Fluid, Schematic
-------
to maximize the efficiency of the cycle, then, are primarily concerned with
practical methods to: (1) raise curve A - E (high boiling and superheat tem-
peratures), (2) lower curve F - A (low condenser temperature) and (3) make
the actual thermodynamic processes in the expander approach the ideal isen-
tropic process described by the vertical line E - F. Secondary efficiency
gains can be obtained in a given system by transferring" internally to other
parts of the cycle any excess heat (which might otherwise be rejected to the
atmosphere) from the condenser or the combustor exhaust. Some systems
also use heat in the combustor exhaust to preheat the incoming combustion
air or use available energy in the expander exhaust to drive parasitic
components.
4.2.2 Component Design Features
Each of the components making up the Rankine cycle engine
system previously discussed perform certain functions and present individ-
ual design problems. A brief discussion of each of these components, includ-
ing some of the more significant and general design approaches, is contained
in this section. For purposes of this discussion, the combustor and vapor
generator will be considered as one component. In most cases they are,
physically, one integrated component.
4.2.2.1 Feed Pump
The Rankine cycle requires relatively little work for the pump-
ing process because this involves the liquid phase only. Feed pump work is
not negligible, however, when vapor generator pressures are high. The de-
sign problem is to maximize the pumping efficiency and reduce the parasitic
work required to drive the pump. This problem is compounded by the wide
range of fluid flow rate required and available input shaft speed resulting
from combined requirements of vehicle power demand and engine speed. The
pump may be required to deliver the same full flow rate against high vapor
generator pressures at both low ("lugging") and high engine speeds. In order
to avoid fouling of the vapor generator heat transfer surfaces it is desirable
to lubricate the pump with working fluid only.
4-7
-------
Current designs involve a variety of piston pumps, including
constant-speed, variable-stroke and variable-speed, fixed stroke designs as
well as other multiple-stage pumps, sliding-vane pumps, and gear pumps.
The parasitic horsepower (hp) requirements run from about 2 to as much as
8 percent of the gross engine hp. Special studies on feed pumps were con-
ducted by Lear Motors and Chandler Evans Control Systems (Ref. 4-4).
4.2.2.2 Combustor/Vapor Generator
Currently three main air pollutants emitted from automotive
combustion systems are of concern: oxides of nitrogen (NO ), carbon
monoxide (CO), and unburned hydrocarbons (HC). It is generally agreed that
NO emissions can be minimized by avoiding long periods of combustion at
X
high temperature in the presence of excess air. Complete oxidation of CO and
HC, however, somewhat requires the opposite conditions. In engine systems
wherein combustion gases also act as the working fluid, maximum cycle
efficiency depends on achieving maximum combustion gas temperatures. By
contrast, peak combustion gas temperatures in a Rankine cycle engine can be
kept at a much lower level, dictated only by the requirements of efficient
combustion and of providing adequate temperature gradient to the vapor gen-
erator. This low combustion temperature can easily be provided over a wide
range of fuel air mixture ratios at atmospheric pressures and with appreciable
dilution of the reactants with chemically inert gases (such as recirculated exhaust
gases), if desired, without strongly affecting cycle efficiency.
The main design problem with the combustor is that of supplying
heat efficiently to the boiler without emitting large quantities of undesirable air
pollutants into the atmosphere. Current combustor designs are of both the
direct-injection type and the prevaporization, premixing type with flame
holders. Most use some degree of exhaust gas recirculation for further
reduction of NO emissions.
x
As with any thermal cycle, there is a loss of efficiency in the
energy rejected to the atmosphere from the hot exhaust gases. The temperature
of the exhaust gases leaving the vapor-generator tube bundle must, of course,
be higher than the inlet temperature of the working fluid into the tube bundle.
4-8
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The latter temperature is essentially that of the condenser discharge
(assuming no regenerator). Allowing for reasonable thermal gradients,
this fixes the lower limit of the exhaust gases, as they exit from the vapor
generator tube bundle, in the range 200 to 250°F. Values in this range are
attained for actual systems at the lower heat-release rates. The exhaust
temperature must rise for higher heat-release rates, however, and can
become 600°F or more at maximum burn rate. Some of this heat may be
retained by transferring it to the blower inlet air, and this is done in some
systems. This produces a gain in overall thermal efficiency at the cost of
increased blower-hp requirement and possibly some combustor design
modifications required to accept the higher inlet air temperature. Attain-
ment of maximum overall efficiency in a Rankine cycle system is dependent
on the careful application of such tradeoffs for each particular system.
As discussed earlier^ high efficiency steam Rankine cycle
engines require a high pressure and high temperature working fluid in the
boiler. The steam boiler design problem, then, is not only one of size and
weight but also one of safety. The general approach for safety taken by most
modern designers of automotive steam engines involves variations of the
"mono-tube" concept which incorporates all of the boiler heat transfer surface
in a single (or at most a few) coils of small diameter tubing. Should a tube
rupture, only a small amount of energy in the vicinity of the rupture can be
involved while the remainder can only bleed down slowly through the long
tubing. These designs are considered explosion proof. Tube ruptures have
occurred in development engines, but none have represented more than an
inconvenience.
Special combustor/vapor generator studies have been conducted
by Solar, Battelle, and General Electric (Refs. 4-4 and 4-5).
4.2.2.3 Expander
The two basic classes of expanders receiving most attention
to date are the reciprocating (piston) type and the turbine wheel. The wide
variety of past and current expander designs result from efforts to make the
4-9
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expansion (work producing) step approach as closely as possible to the ideal
isentropic process indicated in Figure 4-2 as the vertical line E-F.
The major losses in reciprocating expanders are:
a. Mechanical friction of moving parts
b. Heat loss by conduction and radiation
c. Vapor leakage past valves and the piston
d. Incomplete expansion of the vapor
e. Fluid friction or throttling of the vapor
f. Initial condensation of hot vapor in contact with
relatively cold surfaces in the cylinder.
The single most significant potential loss is the last of those
described above, (f). The most common cause of cold surfaces is the cooling
induced by the vapor itself after it has been expanded to low pressures and
temperatures. One way to minimize this surface cooling is to keep the cool,
expanded vapor away from the hot intake end of the cylinder. In the "uniflow"
design concept this is accomplished by arranging the intake and exhaust ports
of a cylinder at opposite ends. Engines of the uniflow type can deliver 73 to
83 percent of the ideal cycle efficiency, while those exhausing the vapor at the
same end as the intake may deliver only 56 to 65 percent. Uniflow expanders
tend to be larger in size, however, and may have greater mechanical friction
losses.
Losses resulting from partial expansion of the vapor in a
cylinder become important in torque considerations. If the full expansion
of high pressure fluid from the vapor generator is taken in a single cylinder,
the average force on the piston will vary widely during the stroke and the
average torque will be low. Horsepower delivered to the wheels can be main-
tained by operating the engine at high rpm, but this creates a need for a
variable-speed transmission. If the intake valves are left open longer, and
full vapor-generator pressure is applied to the piston over a larger fraction of
the power stroke, better torque characteristics are obtained. This "variable
4-10
-------
cutoff" concept may entirely eliminate the need for a transmission. But the
vapor in the cylinder after a late cutoff (closing) of the intake valve cannot
fully expand and a significant partial expansion efficiency loss occurs under
certain conditions. (At part load the efficiency can be very high). With
knowledge of the load cycle, an optimum combination of variable cutoff and
transmission can be selected. Most current automotive reciprocating Rankine
engines make use of some degree of variable cutoff.
Another way to improve torque characteristics and to minimize
condensation losses is the "compound expansion" concept. Here the vapor is
partially expanded in a small cylinder, at high mean pressures, then trans-
ferred to one or more large cylinders for complete expansion. The latter
cylinders have lower mean pressures but may provide the same torque because
of the larger piston areas. In some design concepts the vapor is routed through
the vapor generator between the high and low pressure cylinders for "reheat".
Mechanical friction represents a significant efficiency loss in
reciprocating expanders because of the high piston speeds, pressures and
temperatures and the relatively poor lubricating qualities of most working
fluids. Most automotive reciprocating Rankine engines use some type of
standard lubricating oil in the cylinders. It is generally necessary to sepa-
rate any oil which gets into the working fluid before recycling to the vapor
generator, because it may coat the high temperature tubes and reduce heat
transfer efficiency. In some designs, extensive schemes are employed to
introduce oil into the cylinders, then separate it out before the liquid returns
to the vapor generator. It appears that currently available solutions to this
problem are not fully developed and piston and cylinder wear continues to be
a potential problem. At least one special study of this problem area has been
conducted (Ref. 4-3).
Turbine-type expanders are free of the problems of the fric-
tional wear and lubrication technique, and maximum turbine efficiencies are
of the order of reciprocating expander efficiencies. Turbine efficiency is
particularly speed sensitive, however. Hence, it is strongly affected by the
4-11
-------
widely varying speed and power demands in automotive applications. For
acceptable levels of efficiency, the turbine must be operated at high speeds;
integral speed-reduction gearing and a widely variable transmission are then
required to maintain turbine speed close to that required for maximum effi-
ciency regardless of vehicle wheel speeds. These necessarily high turbine
speeds also lead to problems with turbine shaft vibration and sealing.
4.2.2.4 Regenerator
Regenerators are necessary in those engines using working
fluids which exhaust from the expander in a highly superheated state, usually
at temperatures well above the ambient sink temperature. For a reasonably
efficient thermodynamic cycle and minimization of the heat rejection load on
the condenser, as much of this superheat as possible must be transferred to
the working fluid entering the vapor generator. The basic design
problem in the regenerator is the tradeoff between size and weight of the
regenerator and gains in cycle efficiency. Current regenerator designs
for automotive Rankine cycle engines generally appear adequate, and there is
not a great deal of effort being concentrated in this area.
4. 2. 2. 5 Condenser
Whereas the regenerator's function is to retain heat within the
cycle, the condenser's function is to reject heat to the atmosphere. Generally,
condenser temperature and pressure remain constant as the vapor is condensed.
At a given pressure level, the condenser (or a separate subcooler) cools the
working fluid slightly below the condensing temperature in order to provide
some cavitation margin at the inlet to the feed pump. As with the regenerator,
cycle efficiency strongly depends on cooling the working fluid as closely as
possible to the ambient sink temperature. This leads to large heat transfer-
surface area, size, and weight. With limited frontal area in automotive appli-
cations, a tradeoff of size and weight versus cycle efficiency is required.
One common method of improving the condenser efficiency is forced-air
circulation through the condenser at low vehicle speeds; at high speeds, the
4-12
-------
ram air velocity may be sufficient. Fans are required, particularly for
operation at high-power, low-speed conditions; these fans may require
10 percent of the engine horsepower at full load. Major advances in condenser
design have been accomplished recently by the AiResearch Manufacturing
Co., Division of the Garrett Corporation (Ref. 4-4).
4. 2. 2. 6 Working Fluids
Steam is the most widely used and versatile Rankine cycle
working fluid. At temperature levels required for adequate cycle efficiency,
however, relatively high vapor generator pressures are required. Attempts
to keep boiler pressures low by increasing the degree of superheat yield
diminishing advantages in terms of cycle efficiency. Referring to Figure 4-2,
and the previous discussion on the steam Rankine cycle, it is clear that
maximum work would be obtained from the cycle, for given maximum/
minimum temperatures were the area enclosed by the cycle in Figure 4-2
to be a rectangle. The figure also shows that superheating the steam (line
D - E) raises the peak temperature but, because this process is represented
by a small triangular-shaped area over only the right-hand portion of the cycle,
it does not increase the area bounded by the cycle to any great extent .
Certain organic fluids have thermodynamic properties that
tend to make the Rankine cycle, plotted as in Figure 4-2, look more
"rectangular"; i. e. , more work can be derived from an organic cycle
between the same maximum/minimum temperature limits because saturation
pressures at a given temperature are lower for the organic fluids. These
fluids also normally have a positive slope to the saturated-vapor curve
(rather than negative slope for steam) in the working region (line D - G) on
a Temperature-Entropy plot. This means that little or no superheat is
necessary to avoid condensation in the expander since expansion, rather
than condensing the fluid as in steam systems, vaporizes and superheats
the organic vapor.
4-13
-------
One disadvantage of the organic fluids, however, also results
from this same characteristic. After the desired pressure expansion (usually
to one atmosphere pressure to avoid condenser leakage problems), the organic
fluid is still at fairly high temperature and is still superheated. To reject this
residual heat energy to atmosphere in the condenser would seriously degrade
the cycle efficiency. Thus a regenerator is necessary to transfer excess heat
back within the cycle.
A second disadvantage of organic fluids is the maximum
temperature limitations imposed by thermal degradation. The rather complex
molecules of these organic compounds begin to break down at rather low
temperatures. Thus, while water cycles often operate continuously above
1,000°F, most organic fluids are limited to 600 to 700°F peak temperatures.
Much greater care must be taken in vapor-generator design to avoid "hot
spots" or high temperatures in the fluid film the walls of the tubes.
In addition to thermal stability, many organic working fluids
have factors such as cost, flammability, and toxicity that do not apply to
water. Candidate fluids do exist, however, in which at least the last two
factors appear to be satisfactory. The corrosion factor may be greater or
lesser for the organic fluid than for water. Most organic fluids, however, have
a freezing point well below minimum ambient temperatures. The high freez-
ing point of water is a definite disadvantage for a steam Rankine cycle engine
in automotive applications. As a result of the multiple advantages and dis-
advantages, a wide variety of organic fluids have been investigated, some
of which are considered proprietary by the developers. The two automotive
organic Rankine systems currently under development use different working
fluids. A recent study of optimum organic working fluids has been conducted
by Monsanto Research Corporation, (Ref. 4-4).
4.2.2.7 Transmissions
It seems clear, from past and current automotive Rankine cycle
applications, that all engine systems have a single set of operating conditions
at which the engine delivers the best combination of high efficiency and low
emissions. At other operating conditions the efficiency may fall off severly,
4-14
-------
the emissions may exceed required levels, or both. The off-optimum
operating conditions are forced on the engine system by the widely varying
combinations of wheel speed and power/torque demand of the vehicle. There-
fore, development of an "infinitely" variable transmission would benefit all
types of Rankine cycle engines. This type of transmission functions over a
very large speed ratio (generally stepless) between the engine shaft and the
wheels. The engine could then operate near its best speed at all times. This
is particularly desirable in engines with turbine expanders since, as noted
previously, turbine efficiency is a strong function of turbine speed. Recipro-
cating expanders are less sensitive to engine speed but could also benefit
from this application.
The problems in developing new transmissions are with effic-
iency, size, and weight. Several advanced-design transmission studies have
been conducted recently (Ref. 4-3).
4.2.3 Operating Characteristics
All Rankine cycle automotive engine systems currently under
consideration are designed for minimal effect on normal driving habits
associated with the passenger car. Thus, such vehicles would still be
controlled by operator activation of a conventional accelerator and brake
pedal. In principle, the startup time from cold start need be no longer
than that of a controlled spark ignition engine. However, this area may still
need some more development for vaporizing-type Rankine cycle combustors.
The handling qualities of the Rankine cycle vehicle in most
maneuvers can be expected to be very similar to present-day conventional
automobiles. There will probably be certain conditions, however, in which
a reduced response will be noticeable, such as a sustained power demand
following an extended low-power operation. This results from a certian
minimum thermal inertia inherent in the transfer of heat from one fluid to
another through a wall. Safety considerations dictate that current systems do
not have a very large storage of high-pressure, high-temperature working
fluid that can be drawn upon to meet such transient demands. Short-term
demands, however, for sudden acceleration or for maneuvers such as passing
other cars can be met adequately.
4-15
-------
Internal controls to ensure proper component operation, to
coordinate fluid flows, and to ensure safety are all automatic and can be com-
plex in some designs. There are no fundamental problems in control system
techniques for Rankine cycle vehicles; but the optimum arrangement for a low
cost, reliable control system is a matter to be determined only with the aid
of extensive testing on the engine dynamomenter stand and, subsequently, in
the vehicle.
A key control parameter is the air fuel ratio. The degree
to which the control system can maintain this ratio within prescribed limits
(determined in the steady-state test phase) during all engine transients and
over the full load range will determine the degree to which low emission
levels will be realized in actual vehicle operation.
Air fuel ratio control is, of course, required for any variable-
speed engine, internal or external, and the task is not necessarily more
involved for the Rankine cycle engine as compared to the internal combustion
engine. Rankine cycle engines, however, involve a third fluid, the working
fluid, which must also be controlled in varying proportion to the rate of heat
generation by combustion processes. The problem is compounded by the
fact that power supplied to pump the various fluids is usually taken directly
from the shaft; thus, power is supplied at shaft speeds which are not directly
related to the pumping-power demand.
For example, in a Rankine cycle engine, if a signal for vehicle
acceleration is received, the first control step might be to delay the closing
of the inlet valves in a variable cutoff reciprocating expander or to increase
the steam flow area in a variable admission rotary turbine expander, thus
increasing the demand for steam. A type of "anticipatory" control system
would simultaneously begin to activate those components generating steam.
The combustion air blower, fuel pump, automizer drive (in some cases),
and the working-fluid feed pump would all be called upon to increase flows
and rate of combustion in a coordinated manner, despite the fact that the drive
speed to these components from the engine has not yet begun to increase. As
engine speed increases, the proper flow rates and proportioning of these flows
would have to be controlled despite the increase in drive speed.
4-16
-------
Control subloops in individual components may also be required.
Many vapor-generator designs involve automatic feed-flow proportioning de-
vices that permit bypass of some of the working-fluid flow around the vapor
generator in order to control final vapor temperature. Some sort of fan control
based on condenser outlet temperature may also be required.
4.3 PERFORMANCE CHARACTERISTICS
There are no inherent problems in building Rankine cycle en-
gines in the automotive horsepower .range. A number of Rankine cycle engines
have been built and are being built to fit in the same space and to supply
approximately the same power as existing internal combustion engines.
Although current preprototype automotive Rankine cycle engines
being developed in the AAPS Program have not yet been installed and operated
in vehicles, measurements of emissions and component performance at steady-
state conditions combined with reasonably conservative calculations of system
performance for a car operating over the Federal Driving Cycle (FDC), indi-
cate that exhaust emissions of HC and CO will be about 10 percent of the orig-
inal 1976 EPA standards; NO emissions will probably be on the order of one-
X,
half of 1976 emission standards. Care has to be taken in the starting-sequence
design to avoid excessive HC emission during cold start. Emission measure-
ments have not yet been performed under transient conditions using a fully
automatic control system.
Fuel economy in many of the initial and preprototype designs
was poor; however, current automotive preprototype Rankine cycle engines
appear to be demonstrating fuel economy comparable to prototype internal
combustion engines with 1975 emission controls (Ref. 4-7). Continued
development work on Rankine cycle engines will be guided by the need to
further increase fuel economy while maintaining low emissions.
Significant noise in Rankine cycle automotive engines appears
to emanate from the condenser fans and from the high-speed turbine in rotary
expander engines. The high-frequency turbine "whine" might, in some cases,
represent a higher or more irritating noise level that in the internal combus-
tion engine. Otherwise, these engines are considerably quieter than internal
4-17
-------
combustion engines. External acoustic measurements made during the
California Steam Bus Project showed that noise levels from the Rankine cycle
engine-powered bus were about 10 decibels lower than a diesel-powered bus
(Ref. 4-11).
With the very low level of HC emissions from Rankine cycle
engines, it is unlikely that odor will be a significant factor. In the California
Steam Bus Project using diesel fuels, "only light odors, reminiscent of gas
turbine or jet engine exhaust, were sometimes noted around the steam vehicles."
There is not yet sufficient experience with modern, compact
Rankine cycle automotive engines to assess durability and maintainability.
It remains to be demonstrated that conden_ser designs can function adequately
in all environmental condtions. The large number of thin, closely spaced
aluminum fins necessary for a compact high-performance design might
be susceptible to foreign object damage. The need to reduce condenser
fan parasitic power requires low-restriction, cooling-air inlet designs which
may also fail to restrict entry of foreign objects. On the other hand, an
Australian steam car, using an automotive radiator for a condenser, has been
road tested extensively, and the manufacturer anticipates engine life will
exceed 250, 000 miles (Ref. 4-4).
Primary safety problems (other than those normally associated
with hydrocarbon fuels) are related to the high pressures and temperatures
of the working fluid; particularly, the potential flammability and toxicity of
organic working fluids. Also, any combustion products resulting from burning
of the working fluid by fire is a potential problem.
Uncontrolled escape of the high temperature working fluid
would be an obvious danger, but this is relatively easy to prevent, due
in part to the monotube vapor-generator design previously described. With
careful selection, organic working fluids can also be safe. Both of the or-
ganic fluids being used in the two systems in the AAPS Program are considered
nonflammable and nontoxic, as is the prime candidate selected in the Monsanto
organic working fluid study. The use of water as the working fluid, of course,
eliminates the toxicity problem.
4-18
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4.4 CURRENT STATUS OF TECHNOLOGY
Some of the more significant developments relating to
automotive applications of steam Rankine cycle engines are summarized in
Ref. 4-11. While steam cars preceeded internal combustion engines (ICE),
they were eventually driven out of existence by rapid development of low-cost,
mass production techniques for the ICE. Rankine cycle engines are receiving
renewed interest today because of the recent emphasis on low air-pollution
emissions from mobile cources. No significant number of Rankine cycle
automotive vehicles are in use today other than research vehicles. New
technology in this direction is being fostered by the programs described below.
4.4.1 Current Research and Development
Two complete automotive Rankine cycle system development
programs are currently under way: the EPA AAPS R/D Program and the
California Clean Car Project. The DOT-sponsored California Steam Bus
project was also recently completed. A number of component and working
fluid studies are also current or have recently been completed. There are
also a few steam-powered vehicles of interest under private development.
The system programs will be discussed here first, followed by the compo-
nent and private studies.
4.4.1.1 AAPS R/D Program
The AAPS R/D Program was initiated in 1970 to demonstrate
alternative power systems to the spark-ignition, internal-combustion engine
applicable to full-size passenger cars. Three Rankine system contractors
were selected in 1971, and a fourth was added about a year and a half later.
The four approaches represent the four combinations of water versus organic
working fluids and reciprocating versus turbine expanders. The four
system contractors are:
• Water Working Fluid;
1. Reciprocating Expander. Scientific Energy Systems Corp.
[ (formerly Steam Engine Systems) (SES)], Watertown, Mass,
4-19
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2. Turbine Expander. Lear Motors Corp. (LMC)
Reno, Nevada
• Organic Working Fluid:
3. Reciprocating Expander. Thermo Electron Co.
(TECO), Waltham, Mass.
4. Turbine Expander. Aerojet Liquid Rocket Co.
(ALRC), Sacramento, Ca.
Development of a low-emission preprototype engine system by
each contractor was the primary goal, followed by selection of a single pro-
totype design for installation and testing in a vehicle. The contractors went
through several iterations in component designs during the preprototype devel-
opment phase. Proposed designs for prototype engines are nearing completion.
Figure 4-3 shows a mockup of a typical installation in an engine compartment
for a 6-passenger car. Figures 4-4 through 4-7 are photographs of the four
different preprototype engines; Figures 4-8 through 4-10 indicate the manner
of prototype installation of one of these systems in a conventional passenger
car. Discussions of contractor progress with concepts, designs, design
problems, and tests are contained in the AAPS Contractors Coordination
Meeting Summary reports (Ref. 4-1 through 4-5, 4-7 and 4-8). Table 4-2
lists some of the significant preprototype power plant design features for
each contractor. Proposed prototype designs are different in some details
from those indicated in the table. Comments on the data presented
in Table 4-2 in context with discussions of the previous sections follow.
Three of the four combustor designs involve vaporization of
the fuel and mixing with air prior to entering the combustor, with the
attendant need for flameholders. This technique is capable of producing
very low steady-state emissions of all three pollutants but requires
considerable design attention with respect to low-HC emissions during
cold start. None of these combustor designs uses staged combustion, and
three of the systems use varying degrees of EGR, for NO emission
control.
4-20
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I
rv
Figure 4-3.
1 50-Horsepower Steam Engine (Scientific
Energy Systems Corporation)
-------
Figure 4-4. Steam Reciprocating Engine (Scientific Energy Systems Corporation)
-------
M
U)
Figure 4-5. Organic Reciprocating Engine
(Thermo Electron Company)
-------
Figure 4-6.
Organic Turbine Engine
(Aerojet Liquid Rocket Company)
-------
Figure 4-7. Steam Turbine Engine
(Lear Motors Corporation)
-------
(NJ
STEAM GENERATOR
STARTER
0
A
\
ALTERNATOR
POWER STEERING PUMP
EXPANDER
FEEDPUMP
ll/
AIR CONDITIONING
COMPRESSOR
A
\
\
Figure 4-8. Prototype Installation, Condenser Removed, Plymouth "C" Body
[(Front Interior View) (Scientific Energy Systems)]
-------
-SO
I
t\)
-4
Figure 4-9. Prototype Installation, Plymouth "C" Body, Top Section
(Scientific Energy Systems)
-------
&•
I
00
J- ©
Figure 4-10. Prototype Installation, Plymouth "C" Body, Side Section
(Scientific Energy Systems)
-------
Table 4-2. AAPS Rankine Cycle Engine Program
Preprototype Power Plant Descriptions
Contractor
Working
fluid
Expander
Combust or
Vapor
Generator
Regenerator
Condenser
Expander
Control
Feedpump
SES LMC TECO ALRC
Water
4 -in-line
single-
acting,
trunk-piston
uniflow
Air-
atomizer,
pre-
vaporizing,
EGR
Monotube
None
AiResearch
hydraulic
drive
Variable
cutoff
3 -in -line
piston,
variable
delivery
(direct drive
Water
Single-stage
supersonic
impulse
turbine
Air-
atomizer,
pre-
vaporizing,
pre -heater
EGR
Monotube
Steam-to -
air
Harrison
belt-drive
Variable
admission
Variable -
displace-
ment
Fluorinol-
85
V-4 single-
acting,
trunk-piston
uniflow
Direct-
injection,
r ota ry -
atomizer,
EGR
Monotube
Vapor -to -
liquid
AiResearch
belt-drive
Variable
cutoff
Variable -
displace-
ment,
radial-
piston
AEF-78
(Super-
critical)
Single-stage
supersonic
impulse
turbine
Pre-
vaporizing
pre -heater
Super-
critical,
multi-tube
Vapor -to -
liquid
AiResearch
hydra ulic -
drive
Variable
admission
Variable -
displace-
ment, vane
4-29
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Vapor generators are predominantly of the monotube design.
Both organic working fluid systems use a regenerator; the
water-based systems either use no regenerator (SES) or use excess heat
in the expander exhaust to heat the incoming combustion air in a sort of
hybrid regenerator/preheater (LMC).
All three of the initial contractors made use of the AiResearch
plate and fin condensers supplied them by the government. The LMC con-
denser was manufactured by Harrison Radiator Division of General Motors.
In each case the major engine power control is exercised on
the expander. The reciprocating expander designs use the variable cutoff
concept with the resulting partial expansion losses and favorable torque
characteristics discussed earlier. Turbine power control in both cases
is obtained by varying the steam flow area (admission) to the turbine.
Preprototype feed pumps are generally of the variable-
displacement type. There does not appear to be general agreement on
the best feed pump design. This is also true of the two independent feed
pump studies discussed later.
It appears too early in these current system programs to
tabulate conclusive data for exhaust emissions or fuel economy. Much of
the testing and effort to date has been in component development with low
emissions as the primary design goal. Emissions are measured as con-
centrations of pollutants in the combustor exhaust during steady-state
engine dynamometer tests. To properly extrapolate these data to emission
levels expected from a full-engine system installed in a vehicle driven over
the Federal Driving Cycle (FDC) requires accurate knowledge of: (1) engine
system and power-train efficiencies over the full range of speed and power
required in the FDC, (2) cold-starting, exhaust-emission characteristics
and (3) the effects of engine system transients on emissions. To date, these
data have not been fully established for any of the systems in the program.
Based on the steady state data and what are believed to be reasonable analytical
procedures, the conclusion is that the Federal Driving Cycle emissions will
4-30
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be quite low. These preliminary data are shown in Table 4-3. These
computations may not adequately describe emissions during transient
operation and HC emissions during cold start.
Table 4-3. Preliminary Emission Results on Federal Driving Cycle
(gm/mi), Reference 4-12
Water Reciprocating
Water Turbine
Organic Reciprocating
Organic Turbine
HC
0.09
0. 21
0.02
0. 16
CO
0.50
0.70
0. 19
1.01
NO
X
0. 20
0.38
0.25
0. 13
NOTE: These values are based upon steady- state measurements and
analytically converted to the 1972 Federal Test Procedure.
From the small amount of available fuel economy data, it
appears that preprototype systems have demonstrated equivalent fuel economy
to comparable spark ignition engines with emission controls. These data,
again based on steady-state tests, are shown in Figure 4-11. This plot
compares projected preprototype fuel economy data with preliminary data on
1975 spark-ignition, prototype vehicles. The data have been corrected to the
same vehicle weight and test procedures.
Measured preprototype engine system weights tend to be
several hundred pounds greater than the comparable spark-ignition
engine. Comparable weight was not of prime concern in the preprototype
design. Predicted weight reductions in prototype engines narrow this gap
somewhat. If these predictions are borne out, the small addition to vehicle
weight should not have a major effect on fuel economy and emissions.
4-31
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OJ
20
19
18
.2 17
v. 1<5
-------
The large reduction in frontal area and volume achieved in the
AiResearch condenser, discussed later, has contributed strongly to the
success in automobile underhood packaging of each Rankine engine system
under development. It seems clear, however, that all available space will
be required. It remains to be demonstrated whether the high-density, thin,
delicate plate and fin construction of these condensers is practical in the
rugged usage associated with many typical automotive operations. Demon-
strations must be carried out with a prototype system installed in a car
and field tested.
4.4.1.2 California Clean Car Project
The California Clean Car Project [(CCCP) (Ref. 4-9)] was
initiated in November, 1972 by the California State Assembly. This project
differs from the AAPS Program primarily in that it addresses the compact
car field. The AAPS Program considers a vehicle of about 4,300 pounds
curb weight, while the CCCP requirements call for curb weight in the range
2,300 to 2,700 pounds. Aerojet Liquid Rocket Company (ALRC), Sacramento,
California and Steam Power Systems (SPS), San Diego, California were the
contractors selected. Initial evaluation of both cars is scheduled for summer,
1974. With little more than 18 months for each contractor to supply an
operating vehicle, the project primarily addresses technological feasibility
and low emissions with concern for fuel economy to be undertaken in a later
program.
Both contractors are using water as the working fluid: ALRC
uses a single-stage impulse turbine expander with gearbox as in its AAPS
program engine, but in this case includes an infinitely variable hydrostatic
transmission. The ALRC vehicle uses a monotube vapor generator procured
from Scientific Energy Systems (SES) that is very similar to the SES vapor
generator in the AAPS Program.
The SPS engine uses a four-cylinder, double-acting, compound-
expansion expander with continuously variable cutoff. The steam generator
is being developed by Solar Division of International Harvester. The
4-33
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combustor design is similar to that developed in a special combustor study
for the AAPS Program (discussed later). Engine power and speed control
is accomplished largely through variations in the expander steam cutoff,
but a two-speed transmission is also used for achieving smooth torque and
power changes to the rear wheels.
Both contractors are well into the component test stage and
engine system assembly is about to begin. Preliminary system emissions
and fuel economy data, again based on early steady-state component testing.
appear to be in the same range as data obtained to date in the AAPS Program.
Contractor estimates of exhaust emissions are all well below Federal 1976
standards, and fuel economy is predicted to be equivalent to the levels
demonstrated by comparable spark-ignition engines. As yet no engine
system or installed vehicle system tests have been conducted.
4.4.1.3 California Steam Bus Project
Although this report concentrates on alternative engines and
fuels for automobiles, experience with related automotive projects (in this
case buses) provides valuable data (Ref. 4-11).
The Steam Bus Project was initiated in June, 1970. Three
contractors were selected:
a. William M. Brobeck & Assoc. (Brobeck), Berkeley,
California in conjection with the Alameda Contra Costa
Transit District (A-C) based in Oakland, California.
b. Lear Motors Corp. (LMC), Reno, Nevada, with the
San Francisco Municipal Railway (MUNI).
c. Steam Power Systems (SPS), San Diego, California
with the Southern California Rapid Transit District
(SCRTD) of Los Angeles.
Funding was provided by the California State Assembly under
a Federal research grant from the Urban Mass Transportation Agency.
Overall project management and nontechnical evaluation was conducted by
Scientific Analysis Corporation of San Francisco. International Research
and Technology Corporation, Washington, D.C. provided technical management
4-34
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and evaluation services. Exhaust emissions were evaluated by the California
Air Resources Board (ARE).
The project was completed in September, 1972 following an
effort of about two years. All three of the buses were demonstrated in actual
revenue service accumulating about 8,370 miles of road testing and service
under steam power.
As in the California Clean Car Project, it was recognized at
the outset that available time and funds were quite limited. As a result project
objectives stressed the demonstration of practical, operating standard buses
with low emissions and did not include any fuel economy goals. As a technical
requirement, complete and design-inherent safety received the highest priority.
Standard 40-foot by 8. 5-foot buses, originally configured to seat
51 passengers, were supplied to the contractors. As a time and cost-saving
expedient, existing commercially available transmissions were employed, even
though the torque converters were not well matched to either the reciprocating
or turbine expander engines. For the same reasons, fixed-cutoff reciprocating
expanders with steam throttling control were employed rather than the more
difficult to develop, but more efficient, variable cutoff designs. Thus the gains
in efficiency and torque characteristics obtainable from the optimized tradeoff
of steam cutoff and transmission design were eliminated at the outset. The LMC
system used partial-admission turbine control.
One distinct advantage of this project is that the three develop-
mental Rankine cycle engines were installed in buses and operated over driving
cycles as nearly identical as possible to those currently negotiated by optimized
diesel engines. Thus the comparison between steam and diesel bus operating
characteristics is direct, practical, and conservative. It is perhaps this direct
comparison, rather than any absolute values for these steam engines, that is of
greatest value.
Table 4-4 describes some of the salient design features of
the three steam-bus engines developed. All systems used water as a
working fluid largely because of the strong emphasis on safety and the
4-35
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Table 4-4. California Steam Bus Project,
Power Plant Descriptions
Contractor
Working Fluid
Expander
Combustor
Vapor
Generator
Regenerator
Condenser
Expander
Control
Transmission
Brobeck
Water
3-cyl. double-
acting,
compound -
expansion
Air-atomizer,
tangential -
firing
axial-flow
Modified
monotube
None
Remote
hydraulic
fan drive
NA
Dana-spicer
No. 184,
2-speed
torque con-
verter (direct
above 29 mph)
LMC
Water
Single-stage
supersonic
impulse
turbine
Air -atomizer/
spinning cup
atomizer
radial-flow
Monotube
Steam-to-air
Beltdriven
fans
Variable
admission
Reduction
gearing,
4-speed
automatic
(Allison
HT-740)
SPS
Water
6-cyl. double-
acting,
compound -
expansion
with reheat
Air-atomizer
Series/
parallel-tube
None
NA
Steam-
throttle
GM-Allison
super V
(original bus
transmission)
NA Not available
4-36
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inherent simplicity of water systems. Both contractors employing reciprocating
expanders chose to use double-acting, compound-expansion designs ; and the SPS
system uniquely included reheat of the steam between expansion stages. Some
form of monotube steam-generator design, with its inherent safety, was used
by all three contractors. The LMC engine was initially designed for an organic
working fluid, requiring a regenerator. When the system was changed to water,
the regenerator was retained in the system. Regenerators are not used in the
other two systems.
Selected power plant specifications are shown in Table 4-5.
Condenser frontal areas appear to be 3 to 6 times those of the AAPS Program
Table 4-5. California Steam Bus Project, Selected
Power Plant Specifications (Ref. 4-11)
Contractor
Gross bhp
Rated net system bhp
Steam pressure, psi
Steam temperature, °F
Lowest BSFC, Ib/net bhp-hr
Condenser frontal area, ft
Approximate weights, Ib:
Boiler with burners
Expande r
Condensers with fans
Transmission
Auxiliaries
Other
Total system, dry weight
Brobeck
240
200
1,000
850
0.985
34.5
920
965
750
625
491
1,026
4, 777
LMC(a)
249
180
1,000
1, 000
1. 13
19.4
890
110
420
700(b)
392
590
3, 102
SPS(a)
275
224
1,000
750
1. 18
32. 2
850
1, 250
800
600
800
400
4,700
Derated from figures shown for use in bus system.
*b)With gearbox.
4-37
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automotive condensers, although the bus-engine gross horsepower is less than
twice that of the automotive engines. This reflects the improved AiResearch
condenser designs in the AAPS Program. Engine system parasitic losses as
fractions of gross horsepower appear to be in the same range as those in AAPS
Program automotive engines. Only one of the bus-system designs (Brobeck)
could be incorporated in available space without requiring a. reduction in seat-
ing capacity. Both of the other buses had five less seats (nearly 10 percent
reduction) than the original diesel-powered versions. However, the Brobeck
system used the existing engine compartment only for the steam generator and
its automatic controls; all other system components were located under the
floor.
A comparison of average emissions of the three steam bus
systems with those measured from four diesel powered buses is shown in
Table 4-6.
Table 4-6. Comparison of Emissions from Steam- and Diesel-
Powered Buses (Ref. 4-11).
Steam Buses
Diesel Buses
Calif. Heavy -Duty
Engine Standards, 1975
Emissions, g/bhp-hr
CO
4.0
4.3
25
HC
0.9
1.4
-
NOX
1.8
10.4
-
HC+NO2
2.7
11.8
5
4-38
-------
These emissions data were obtained on the complete, installed engine
systems, using a chasis dynamometer procedure over a standard
driving cycle. On the average, all steam-bus emissions were less than
those of the diesel buses. The average steam-bus NOX emissions were
only 17 percent of those from an average diesel bus and the HC emissions
were 64 percent of the diesel. The average CO emissions from the two
reciprocating expander-powered steam buses were 63 percent of the average
for the diesel buses. (The LMC system tended to emit 2 to 3 times the CO
of the other two Rankine systems). Thus, while the diesel engine is currently
considered relatively "clean" in terms of CO and HC emissions, the steam
systems tended to emit about two-thirds of these emissions. Also, the
emissions shown above are in terms of grams of pollutant per brake
horsepower-hour. The low steam bus emissions were obtained, then, in
spite of the much larger fuel consumption of the steam engines for the same
power supplied.
Fuel consumption of steam buses was 3 to 4 times higher than
for the diesel buses. Under actual stop-and-go, bus-route driving conditions
the steam buses obtained 0. 52 to 1. 1 mpg while the diesel buses delivered
2.3 to 3.7 mpg. Idle fuel consumption of the steam buses was more than
6 times that of the diesels. Some of the reasons for the low fuel economy
of the systems demonstrated in this program were built into the program
from the start, as discussed earlier regarding component design compromises.
Other improvements could be expected in the normal course of extended
development of this type of engine, with fuel economy a stated, high-priority
goal. Certainly diesel engines, used as a standard of comparison in this
project, are in a much more advanced state of development than are automotive
steam engines.
4.4.1.4 Recent Component Technology Programs
Recognizing certain problem areas in the development of
Rankine cycle automotive engines, special component studies have been
conducted as part of the AAPS Program. Some of the more significant
4-39
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studies are concerned with problems of working fluid selection and optimum
design of vapor-generator feedpumps, combustors, condensors, and trans-
missions. These are discussed briefly in the following pages.
4.4.1.4.1 Working Fluids
At least one special study of organic working fluid selection
recently has been completed by Monsanto Research Corporation (Ref. 4-1
through 4-4). (To a lesser extent, organic working fluids have also been
studied by a wide range of other contractors. ) The objective was to identify
superior working fluids for use in Rankine cycle automotive power systems
using positive-displacement and turbine expanders. Selection criteria were
performance, safety, economy, and acceptance (the latter category including
all other important criteria). An initial goal of the study was that the cost of
the initial fill plus replacement charges in a 5-year period should not exceed
$100. Two final candidate fluids were identified:
Mol % Wt. %
a. Pentafluorobenzene 60.0 57-5
Hexafluorobenzene 40.0 42.5
b. Water 65.0 26.4
2-Methylpyridine 35.0 73.6
These two candidates were extensively characterized physically, chemically,
thermodynamically, functionally, and economically.
Candidate (a) satisfies efficiency, temperature, pressure and
density limitations, and regenerator size-limit criteria for both reciprocating
and turbine-expander engines. It is considered fire-safe and unlikely to
cause serious injury on accidental human exposure. It is stable up to
at least 750 F. If allowed to escape to the atmosphere it degrades at a
relatively slow rate. Synthetic lubricants are necessary with this fluid
and several low-cost synthetics show promise. The fluid is compatible with
low-cost, available construction materials.
4-40
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Candidate (b) also satisfied all performance goals except the
goal of 30-percent efficiency for an ideal cycle. It is flammable but less
so than common fuels and is not likely to produce dangerous concentrations
of toxic products in a fire. Accidental, external contact with the fluid is
\
not likely to cause serious injury, although it is more toxic than candidate (b).
It has an extremely disagreeable odor, which is viewed as a desirable charac-
teristic in terms of positive indication of a system leak and avoidance of
voluntary inhalation. It is stable up to 720°F. Metal corrosion poses a
potential materials compatibility problem. The fluid, once released from
the engine, would likely be rapidly decomposed by microorganisms and
would not constitute an atmospheric pollutant. Compatible low-cost
synthetic lubricants again appear necessary and promising.
Neither fluid appears likely to meet the cost goal, however.
For candidate (a), cost to the owner of the Rankine cycle car for a 5-year
supply of working fluid is estimated by Monsanto to be in the $120 to $200
range. The estimated cost of fluorinated benzenes is very sensitive to the
assumed price of liquid fluorine. The 5-year cost for candidate (b) is
estimated to be in the $80 to $150 range. Reasons for the wide range in the
latter estimate were not stated in the references cited.
4.4.1.4.2 Feed Pumps
Each of the automotive Rankine cycle engine system contractors
expended considerable effort on selection and development of working fluid
feed pumps for their systems. Special studies were conducted by LMC and
Chandler Evans Control Systems (CECO), Division of Colt Industries
(Refs. 4-3 and 4-4). Pumps were designed to match requirements of the
engine systems under development in the AAPS Program (both water and
organic working fluids).
The separate pumps selected by LMC to satisfy all require-
ments of the engine systems were of the radial-piston type with variable
4-41
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capacity or variable speed and direct drive. Separation of the working fluid
from the pump oil is provided. An externally-driven boost pump is necessary
under certain operating conditions.
Vane-type pumps were selected by CECO as the best type for
this application. The organic fluid pumps were two-stage, 6.25:1 flowrate
turndown ratio, variable-speed drive. A fixed-displacement water pump
employs a bypass. All pumps are lubricated by the working fluid. For a
more near-term application, gear pumps are preferred because of the lesser
development required.
All contractor designs met the requirements for overall
efficiencies greater than 50 percent at 10 percent of maximum flow and
70 percent between 30 and 80 percent of maximum flow.
4.4.1.4.3 Combustors
Solar Division of International Harvester (Solar) initiated a
combustor/vapor generator development program at the outset of the
AAPS automotive Rankine cycle engine program (Refs 4-1 through 4-4).
Combustor/vapor generator subsystems suitable for integration in the
AAPS engine systems were to be constructed. Separate combustor studies
were also conducted by Battelle-Columbus Laboratories (Refs 4-1 through 4-4)
and, later in the AAPS Program, by the General Electric Co. (GE) (Refs. 4-3
through 4-5).
The Solar combustor involves a spinning-cup fuel atomizer
injecting directly into the primary combustion space, with staged air
admission. The air flow is split approximately equally between the primary
and secondary combustion zones. The final vapor generator design was
of the monotube-type with external fins on the preheater and superheater
(dryer) tubes; an earlier parallel-tube design was subject to flow instabilities.
Based on bench-level tests, all emissions and performance-level requirements
were met except for NOx at the very highest flow rates. No EGR was used.
4-42
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Battelle investigated a wide variety of combustor designs
involving primary zone cooling, staged-air admission (fuel-rich and fuel-
lean primary) and EGR, all with an air atomizer producing "ultra-fine"
fuel atomization. In general, the cooled, rich-primary configuration was
predicted to just meet NOx goals, and the addition of EGR was predicted to
reduce NO well below these goals. Forecast HC and CO emissions were
Ji.
well within required limits.
The GE combustor program was initiated late in the AAPS
Rankine cycle engine system program and was intended to investigate the
feasibility of the use of a nonadiabatic, porous-plate combustor system
for advanced automotive engines. Combustor firings and emissions
measurements were obtained; satisfactory NO and CO emissions were
.X
obtained over a narrow range of loads. HC emissions were very low.
4.4.1.4.4 Condensers
A special study on compact condensers was initiated by
AiResearch Manufacturing Co. Division of Garret Corp. in June, 1970 (Refs.
4-2 thru 4-4, also 4-6). The goal was to develop heat-transfer surfaces and
design techniques that would lead to an automotive Rankine cycle engine-
condenser design having frontal-area, volume, and fan-power requirements
not appreciably greater than those of present automobile radiators. In
the initial experimental-design study, a perforated-fin configuration was
developed which was only 3 percent greater than the frontal area goal and
2 percent greater than the volume goal. The design is of brazed, all-
aluminum construction, uses 22 fins/inch, and rejects about 1.5 x 10
Btu/hour. It can be mass produced at an estimated manufacturing cost of
about $100 including materials, direct labor, and factory burden.
A follow-on development effort was conducted in which con-
denser/fan units were supplied to the three AAPS engine system contractors
(using water and organic working fluids). These units have provided satis-
factory service in the preprototype system test.
4-43
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4.4.1.4.5 Transmissions
Sundstrand and Mechanical Technology, Incorporated,
conducted two short studies on transmissions for advanced automotive engines
(Ref. 4-3). Objectives of these studies were to compare technical and
economic feasibility of transmission types; recommend a transmission for
Rankine cycle automotive applications; and determine performance, cost,
and characteristics of the transmission. All transmission candidates were
of the infinitely variable type. Very early in the study Sundstrand rejected
hydrostatic (poor efficiency), electric (poor efficiency, size, and weight),
and hydrokinetic (no ratio control) systems. Hydromechanical transmissions
were considered attractive for the near term, and traction type transmissions
proved attractive with further development.
4.4. 1. 5 Other Efforts
There are a number of steam-powered vehicles in operation
and under development. This section describes some of the more significant
efforts under way both in the U.S. and abroad.
4.4.1.5.1 The Dallas Project
Under a grant from the Department of Transportation, Dallas
Transit System subcontracted to Ling Temco Vought Aircraft Corporation,
with a further subcontract to Sundstrand Corporation, to build and install
an organic Rankine cycle engine in a small bus (Ref. 4-3). This bus under-
went operational tests, emission measurements (before and after installation
of the new engine), and limited operation on the transit system for evaluation.
The specific objective of the program was to demonstrate operational
feasibility of a turbine-driven, organic Rankine cycle engine in a 25-passenger
bus.
4-44
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The test vehicle was a 25-passenger, TC-25 transit coach
originally powered by a 413 cu. inch Chrysler, V-8 gasoline engine with a
Chrysler 3-speed automatic transmission.
The design criteria for the rankine cycle engine are:
• Supercritical cycle using CP-25 as a working fluid
• Same power-to-weight ratio as full-size city buses
• 700 F peak system temperature
• 235 F condenser temperature
• LPG fuel
• Nonprototypal baseline engine
• ASME Pressure Vessel Code where applicable
• Hermetic CP-25 loop.
The supercritical monotube heater uses a tangentially
induced vortex combustor. The whole heater/combustor is 25 inches in
diameter and 30 in. long including the combustor. Core weight is 150 pounds;
total weight is 375 pounds. The single-stage, partial admissions, impulse
turbine wheel is 6 inches in diameter, rated at 120 shp at 35, 000 rpm
(8,000 rpm idle) with a design point efficiency of 76.6 percent. The
feedpump is a double-acting, variable-speed, fixed-displacement, axial-
piston type, hydraulically driven. The regenerator is a finned tube
cross-counterflow configuration with 12 aluminum fins/inch on copper
alloy tubing. It is 27 inches long by 9. 5 by 12 inches wide with a core
weight of 120 pounds. The effectiveness is 0.90. The condenser is an air
cooled, downflow configuration with aluminum fins on brazed-copper
tubing. (Size: 4 ft. x 4 ft. x 5 in. thick). It is capable of 1, 200, 000 Btu
per hour at 100°F air-inlet temperature. The 6.3-hp, axial-shrouded
condenser fan is hydraulically driven (air flow 16, 800 SCFM). Core
and hotwell weight is 245 pounds.
4-45
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4.4.1.5.2 Carter Enterprises. Incorporated
The Carter Steam Automobile engine has been under
development for about 5 years. Approximately 3 years were spent on
component development, which included four engine and five boiler con-
figurations. After about 10 months on the test stand the current package
was installed in the engine compartment of a standard Volkswagen auto-
mobile and driven for the first time on March 15, 1972. Engine description
and characteristics are:
• Four cylinder, radial, 35 inch with fixed 2 percent cutoff,
6 percent clearance and 14:1 expansion ratio
• 70 hp at 5,000 rpm (same rated speed as standard VW engine)
• 2,000 psi and 1,000°F rated steam conditions at throttle (unit
has operated up to 1, 100 F)
• Standard VW clutch and 4-speed transmission
• Total engine package weighs about 120 pounds more than
the standard VW engine package
• About 30 to 40 seconds to drive off from a cold start
• An early fuel rate measurement showed 17. 8 mpg at
55 to 60 mph. This was before finned tubes were incor-
porated in the boiler, and insulation was added to the
cylinders and valves
• Start sequence is automatic with turn of ignition key;
turndown ratio is 50:1; control is based on pressure
drop across throttle valve
Status and plans for this development are:
• The prototype vehicle has been driven about 2, 000 miles
• Performance and characteristics are being verified by third
parties. Demonstrations will continue.
• Patent applications have been submitted for all appropriate
parts, pieces, and systems of the engine.
4.4.1.5.3 Kinetics. Inc., Sarasota, Florida
This company has built an automobile utilizing an organic
Rankine cycle engine system. Some funding for this work reportedly has been
provided by the Datsun Company of Japan. The Kinetics system utilized a
4-46
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low-temperature working fluid (Refrigerant 113) and a Gerotor-type expander.
The vehicle has no transmission. It is understood that three vehicles have
been built and tested. Emissions or fuel economy data have not been
published.
4.4.1.5.4 Williams Engine Company, Ambler, Pennsylvania
The Williams system employs a monotube steam generator
and superheater with an oil-fired burner. Steam is delivered (through a
throttle valve) to a four-cylinder reciprocating engine.
In its startup mode, patented auxiliary exhaust and regulating
valves lower the compression pressure to permit maximum torque, with
steam injected during five-eighths of the power stroke and expanded during
the balance. However, in its "drive: and "economy" modes - equivalent
to the higher ranges in a customary transmission - the injection of
steam is cut off earlier in the cycle. Then, on the up stroke, a measured
amount of residual steam is recompressed during which process its
temperature becomes higher than that of the steam supply.
Early cutoff of the steam injection is accomplished by moving
the valve camshafts so as to position different sets of cams under the valve
lifters.
To complete the system, exhaust steam is passed through a
water reheater and into a condenser of essentially full conversion into
water. A pump then delivers water at high pressure to the steam generator
as required.
Because of the variable cutoff feature, no transmission is
required and the expander is coupled directly to the driveshaft. The cutoff
selector replaces the conventional gear shifting mechanism, and the throttle
valve functions in the same way as the customary foot accelerator.
Appropriate sensors and automatic sensors and automatic controls regulate
the generation of steam and safeguard the system.
4-47
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In 1966 tests at Mobile Laboratories, the following emissions
data were determined using the California 7-mode cycle test (Ref. 4-13):
Hydrocarbons
Carbon Monoxide
Oxides of Nitrogen
1968
National
Standards
275 ppm
1.5%
1, 500 ppm
Williams Steamer
43 CID Engine (no
anti -pollution devices)
As Measured
20 ppm
0.05%
40 ppm
Corrected to otoicmometric
Air/Fuel Ratio
33 ppm
0.08%
116 ppm
These measurements played a significant role in the renewal
of interest in Rankine cycle automotive power plants, as they were signifi-
cantly below levels of emission from contemporary internal combustion engines.
4.4.1.5.5 Other Developments
A number of other private groups are involved in some form
of development leading toward a Rankine cycle powered vehicle. These are
largely in the component development area, and some are briefly described
here.
The Gibbs-Hosick Trust of Winston-Salem, N. C. has built a
demonstration model of what they call an "Elliptocline" expander. The design
is covered by U.S. Patent 3, 370, 511, and is essentially a variant of a barrel
expander. They have a working model, but no test data to report.
Closed Cycle Systems, Maspeth, New York, is working on system
incorporation of a steam turbine utilizing a pressed-steel disc for the blades.
D-Cycle Power Systems, Richmond, Virginia, is working on a
cycle claimed to be novel, in which steam is recompressed to minimize heat
addition. It is understood that they are building a system for test.
Horst Power Systems, Saratoga, California, has built a model
of a rotary positive displacement expander. This has been run on steam in
their laboratory. No data are available.
4.4.1.5.6 Foreign Programs
A number of programs to develop Rankine cycle engines are
known to be under way outside of the U .S. Unfortunately: detailed informa-
tion is nut available at this time except for the activity described here.
4-48
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Pritchard Steam Power Proprietary Ltd. , Melbourne,
Australia. Steam power systems have been in various stages of development
in Australia over the last 20 years. An engine was developed and operated in a
5 ton truck in the early 1950s.
Extensive road tests were carried out on the Pritchard Engine
in a 1963 Ford Falcon, in November, 1971, in Melbourne. Over 1,300 miles
were run with an identical comparison car that was conventionally powered.
Selected routes went through congested traffic areas as well as on freeways.
Tests included acceleration, hill climbing, starting, noise, and emissions.
The fuel was kerosene. Mileage over a varied 58.4-mile city course was
15.3 mpg. Measurements of NO showed 0.595 gms/mi. Performance
characteristics were directly competitive with or better than the comparison
car.
The system includes a monotube boiler, a uniflow piston
expander with variable valve cutoff, and an exhaust-driven motor which
drives the feedpump and the condenser fan. The condenser is an auto-
motive radiator. The fuel pump and combustion air compressor are
electrically driven. Roller bearings are used on the crankshaft and the large
end of the connecting rods. Engine life is expected to exceed 250, 000 miles.
First cost is projected to be the same as the conventional engines; life time
cost should be less.
SAAB-SCANIA. SAAB-SCANIA, Aerospace Division, Stockholm,
Sweden, has a program which has been under way since 1968 (Ref. 4-10).
Computer programs were initially developed to model engine characteristics,
followed by an evolution of preliminary system designs and, lastly, by com-
ponent development, which is now under way. Present hardware is being
used as a development tool.
Auxiliaries and accessories are driven at constant speed by
a single-cylinder auxiliary expander. Superheated steam at 1500 psi and
600 to 775°F is used. The vapor generator consists of a parallel arrange-
ment of a large number of very small-diameter tubes in intimate thermal
contact. These very closely spaced tubes permit laminar flow of the
4-49
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combustion products through the tube bundle. Vapor generator output
temperature is controlled by a distribution valve between a single preheating
section and two parallel boiling and superheating sections. The main
expander is a 9-cylinder wobble plate device operating at 3000 rpm and
directly connected to the differential. Power control and engine reversal
are accomplished with a variable cutoff valve. The overall concept of the
propulsion system eliminates the need for a transmission.
The condenser consists of two heat exchangers, one an air-
cooled condenser sized for road-load conditions and a "condenser buffer"
for transients. The condenser uses 80 fins/inch on both air and steam
sides, with laminar flow on the air side. An "accordion" design is used
to increase the effective frontal area.
SAAB-SCANIA indicates that the engine can develop 160 hp
continuously and 250 peak hp for short periods. The total engine weight
and fuel consumption are less than those of a conventional 100-hp engine,
and space requirements are about the same. Emissions are considered
"potentially low", and noise and vibration are considered "almost non-
existent". The engine is still considered experimental.
Italian National Propulsion and Energetic Center. This
center, located at the Politecnico in Milan, is performing research centered
around a 100-hp bus engine. The engine will incorporate an organic work-
ing fluid, a high-efficiency multistage turbine expander, and an advanced
regenerator. Emphasis in the design is to achieve as high efficiency as
possible. This will, among other factors, reduce the waste heat and thus
the condenser size. The use of regenerative braking is also being evaluated,
with both mechanical and electrical methods being considered. A detailed
user survey is presently in process. Although centered on the bus applica-
tions, the design can be scaled to passenger-car application.
4.5 PROJECTED STATUS
If it is assumed that the major determining factor in the
emergence of a new type of automotive engine will be the need for simultaneous
lo\v pollution emissions and low fuel consumption, it then appears
4-50
-------
that the Rankine cycle engine represents a viable candidate. Current data
indicate that emissions can be controlled to very low levels without strong
effects on fuel consumption. These same data combined with powertrain
analyses for an automobile indicate that early, preprototype systems could
offer fuel economy comparable to 1975 prototype spark ignition automotive
engines. Improvements currently being tested in the more advanced Rankine
cycle automotive engines described in this report indicate that significant
gains in fuel economy may be obtained. However, these gains have not
yet been demonstrated in systems installed in vehicles. Further, the
ultimate fuel economy obtainable after extensive development is not clear.
The Rankine thermodynamic cycle, within practical limitations of pressure and
temperature, is not one of inherently high efficiency compared to, for example,
Otto or Diesel cycles. Current projections of fuel economy over the FDC for
the early AAPS Program prototype systems are about 10 to 15 mpg. It is
estimated that fully developed Rankine cycle automotive engines probably
could provide very low pollution emissions with fuel economy approaching
12 to 20 mpg in full-sized cars over the FDC. The viability of the concept
may well depend on the changing balance between emission and fuel consump-
tion requirements and on the success of internal combustion engine designs
in meeting these requirements.
From the standpoint of energy conservation and allocation
in general, external combustion engines may be more attractive in the future
because of the wider range of fuels that can be satisfactorily used in them.
Efforts to conserve energy and to develop new energy sources, as shortages
become more critical, may result in the availability of larger supplies of
lower-grade fossil or synthetic fuels. External combustion engines may
be the only automotive power plants that can operate efficiently with these
fuels. Multifuel capability would probably be associated with an increase
in emissions (particularly NO ) over those reported herein for the AAPS
X
Program, since the AAPS-developed combustors are predominantly of the
vaporizing type. They consequently require a fuel volatility similar to that
of gasoline.
4-51
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The low power density of a Rankine cycle engine has not been
completely eliminated in the present designs; thus, these engines are larger
in size than other competing engine systems. Careful attention to packaging,
however, should minimize the impact of engine size on vehicle aerodynamic
drag and energy dissipation. It appears premature to assess the mass pro-
duction cost of the Rankine cycle engine as the internal combustion engine,
since the necessary manufacturing processes have not yet been defined.
In summary, then, the automotive Rankine cycle engine is
still a leading candidate as a viable alternative to the spark ignition, internal
combustion engine. Continued development efforts and supporting research
are planned as long as there is evidence of satisfactory progress in improv-
ing system performance and approaching original program goals of low
exhaust emissions without sacrificing competitive fuel economy levels.
4-52
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SECTION 5
-------
5. STIRLING ENGINES
5. 1 INTRODUCTION
5.1.1 General Description
The Stirling engine is an external combustion, closed cycle,
piston-type power plant that uses a gaseous internal working fluid, usually
hydrogen or helium (see Figures 5-1 and 5-2). Cyclical heating and cooling
varies the pressure of the fluid within a closed volume; this pressure acts
upon a piston, thereby developing output power. The heat transfer processes
are accomplished at high efficiency by alternately and regeneratively moving
the working fluid between heater and cooler sections. Thus, the engine con-
tains at a minimum one expansion volume, one compression volume, one
regenerator, one heater, and one cooler.
The ideal Stirling engine efficiency is 100 percent of the Carnot
efficiency obtained between peak and minimum fluid temperatures. Various
losses reduce the actual efficiency to about one-half of the ideal value.
Nevertheless, the Stirling engine is one of the most efficient prime movers
known.
5.1.Z Historical Background
The Stirling-cycle engine is not a recent development, having
been patented in 1816 by Robert Stirling, a Scottish clergyman. Thousands of
engines were built in the nineteenth century because they were substantially
safer than existing steam engines, although slow and inefficient. They were
eventually made obsolete by the development of the internal combustion
engine and improvements in steam-engine boilers.
In 1938, N.V. Philips Laboratories at Eindhoven, Holland
undertook the development of a modern version of the Stirling engine that
could operate on a variety of fuels, originally as a means of generating
electric power in remote areas. Although the market for such generators
became very limited, continued research resulted in development of the
rhombic drive displacer engine in 1953 and the roll-sock seal in I960.
This configuration had both the power piston and displacer piston in a single
5-1
-------
HEATER
REGENERATOR
COOLER
0 PHASE
ANGLE
DISPLACER
PISTON
POWER
PISTON
SYNCHRONIZER
Figure 5-1. Basic Stirling Displacer Engine, Schematic
DISPLACER
PISTON
POWER
PISTON
Figure 5-2. Stirling Displacer Engine with Rhombic
Drive, Schematic
5-2
-------
cylinder. The rhombic drive made straightline rod movement possible while
retaining the proper phase position between the pistons, allowed small-
diameter seals to be used, and opened the way to high working pressure and
high power. However, the most significant advancement in reducing engine
volume and complexity occurred in 1970 at Philips with the successful opera-.
tion of a double-acting Stirling engine with a swash-plate drive mechanism.
The engine consists of four separate, interconnected cylinders in which the
piston motion is phased at 90-degree intervals by the action of the swash plate.
Each piston serves alternately as a power piston for one cylinder and as a
displacer piston for the adjacent cylinder. This design represented a signifi-
cant improvement in engine volume and complexity over other double-acting
configurations.
5.2 POWER PLANT DESCRIPTION
The Stirling engine is a closed-cycle, external-combustion
engine producing power by expanding a compressed volume of any type of
heated gas. The Stirling cycle has a very high theoretical efficiency equaling
Carnot cycle efficiency; actual efficiencies are about one-half of the theoreti-
cal values. The basic problems encountered in the development of this engine
have been high weight, high volume, and packaging. But improvements have
been made in recent years in reducing weight and volume by the use of double-
acting pistons, which cut in half the number of pistons required (compared to
the displacer engine), elimination the rhombic-drive gearing, and by the use
of high pressure (200 atm) gaseous hydrogen as the working medium (Ref. 5-1).
The thermal efficiency of the engine is improved in inverse relationship to the
molecular weight of the working gas; low molecular weight has a further effect
in reducing engine size.
A thermodynamic description of the Stirling cycle is given by
Figure 5-3, which shows that the theoretical Stirling cycle is a closed regen-
erative cycle (heat stored and recovered) composed of two isothermal and
two constant-volume processes with the regenerative action occurring during
the constant-volume processes. The pressure-volume diagram shows a com-
pression process from point (a) to point (b) where the piston does work
5-3
-------
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HEAT IN
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STORAGE
HEAT
STORED
VOLUME, V
HEAT
-TAKEN
FROM
STORAGE
HEAT
STORED
ENTROPY, S
Figure 5-3. Thermodynamic Description of the Stirling Cycle
(Ref. 5-2)
5-4
-------
compressing the gaseous medium at constant temperature as the cylinder
volume decreases from its maximum value at point (a) to its minimum value
at point (b). During the period (b) to (c) when the working-fluid volume is
held constant, heat is transferred from an energy storage reservoir to the
working medium thereby raising its temperature and pressure to maximum
values at point (c). At this point the expansion stroke begins and continues
to point (d). During this stroke the working medium does work on the piston
at constant temperature as the cylinder volume increases from its minimum
value at point (c) to its maximum value at point (d). At the end of the expan-
sion stroke the working fluid is held at constant volume while heat is trans-
ferred to the storage reservoir. During this process the working fluid
temperature and pressure drop to minimum values.
If regeneration is 100 percent efficient, the amount of heat
stored during the constant volume process following expansion is equal to
the heat taken from storage during the constant volume process after com-
pression. Also, an external combustion process is utilized to transfer heat
during the expansion process to maintain a constant temperature T, . During
the compression process a radiator or heat exchanger transfers heat out into
the environment to maintain the working medium at a constant temperature T_.
5.2.1 Power Plant Configuration
The basic components of the early Stirling displacer engine are
shown in Figure 5-1 and consist of a displacer piston, power piston, heater,
cooler, and regenerator. The motion of the displacer piston is synchronized with
the suction of the power piston so as to produce the required pressure/volume
relationships for the Stirling cycle in conjunction with periodic cooling and
heating of the gas by the heater, cooler, and regenerator.
The rhombic drive, illustration in Figure 5-2, was a develop-
ment accomplished in 1953 which provided a means for actuating the displacer
piston in proper phase relation with the power piston, both operating in single
cylinder. An additional feature of the rhombic drive unit is that it contains a
buffer space which allows the crankcase to remain at atmospheric pressure.
Mass motions in the design result in perfect dynamic balance of a one-cylinder
engine.
5-5
-------
Figure 5-4 shows the efficiency of a 225-hp per cylinder
rhombic drive engine with air, helium and hydrogen as alternative gaseous
working mediums. As shown, high power levels combined with high effi-
ciencies (of the order of 40 percent) are achievable (particularly true for
the hydrogen working medium). This was the most common form of Stirling
engine design under development between the years 1953 and 1971 by the
Philips Company and its licensees, General Motors Corporation, United
Stirling of Sweden, and M.A.N./MWM of Germany. As shown in schematic
form in Figure 5-5, the rhombic-drive, displacer-type engine has single
acting pistons arranged in pairs and linked together through a common
drive mechanism. One is called a displacer piston and the other is called
a power piston. The rhombic drive consists of two meshed timing gears,
two crankshafts, two power-piston connecting rods, and a displacer-piston
yoke. Heat is added by an external combustor to the heater tubes connected
to the cylinder head on one side of the displacer piston. The water-cooled
cooler tubes are connected to the other side of the cylinder between the
displacer piston and the power piston. The heater tubes and cooler tubes
are connected to each other through the regenerator, which usually consists
of a small volume filled with a fine-wire matrix. At any instant the working-
fluid pressure is essentially uniform, since the working fluid spaces (formed
by the two pistons) are connected to each other through the cooler tubes,
regenerator, and heater tubes.
In operation, the motion of the power piston lags the motion
of the displacer piston by about one quarter of a revolution. When the power
piston is in the downward position and moving very slowly, the displacer
piston is moving upward and rather rapidly. This forces the bulk of the
working fluid around through the heater tubes, regenerator, and cooler
tubes to the cold space between the displacer and power pistons. In so
doing, the bulk of the fluid is cooled as heat is transferred to the regenerator
and the fluid pressure drops. The power piston then moves upward, com-
pressing the fluid, increasing its pressure, and transferring heat from the
bulk of fluid to the cooler tubes. The displacer then moves downward, moving
5-6
-------
60%,
O
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50
40
30
250
1000
1250
AIR
20
10
225 hp/CYLINDER
HEATER, 700°C
COOLING WATER, 25°C
GAS PRESSURE 110 kgf/cm*
I
I
20 40 60
SPECIFIC POWER OUTPUT, horsepower/liter
Figure 5-4. Stirling Engine Efficiency versus
Specific Power Output (Ref. 5-3)
-------
HOT EXPANSION
SPACE
HEATER
DISPLACER PISTON
REGENERATOR
COOLER
COLD COMPRESSION
SPACE
POWER PISTON
EXTERNAL SEALS
SYNCHRONIZING
GEARS
RHOMBIC DRIVE
(Enables straight-line
movement of the rods so that
external seals of small diameter
can be used.)
Figure 5-5. Displacer Type Engine, Schematic (Ref. 5-3)
5-8
-------
the bulk of the fluid back through the regenerator where heat is transferred,
thereby further increasing its pressure. The power piston then moves
downward as the work fluid expands and moves the bulk of fluid through
the heater tubes, thereby effecting a heat transfer and completing the cycle.
These motions do not occur separately but are brought about by the approxi-
mately 90-degree, out-of-phase motion of the two pistons.
Because the internal fluid pressure is low when the power
piston moves upward and high when the piston moves downward, a net
work output is produced. The effect of the regenerator is to improve
efficiency by storing heat from the hot fluid as it flows toward the cold space
and by releasing the heat when the flow reverses.
Fuel is burned in the combustion chamber in the vicinity of
the heat transfer tubes, and the combustion gases pass through this array of
tubes surrounding the combustion chamber. The exhaust gases leave at high
temperature and pass through the preheater, which cools the exhaust and
warms the incoming combustion air. This technique conserves heat that
would otherwise be wasted.
The most significant advancement in reducing engine cost and
volume has come about through application of the double-acting Stirling engine
concept and swash-plate drive mechanism. An engine of this design was run
at the Philips Laboratories in 1970 (Ref. 5-3). Shortly thereafter, the Ford
Motor Car Company entered into an agreement with Philips to develop a
swash-plate drive, four-cylinder, double-acting Stirling engine for automo-
bile application. (See Figures 5-6 through 5-9.) A picture of the engine
mockup for a Ford Torino automobile is shown in Figure 5-9. Preheater
and burner are at the front. Exhaust pipes show at bottom, with an exhaust
gas recirculation tube extending upward to an intake-air blower. A combined
water pump and power steering pump, belt-driven, is located near the front,
and the alternator is shown at the rear corner.
United Stirling of Sweden is also developing a four-cylinder,
double-acting Stirling Engine with a conventional crankshaft for another
5-9
-------
Ul
I
Figure 5-6. 1973 Ford Torino with Philips Stirling
Double-Acting Engine (Ref. 5-13)
-------
REGENERATOR
COOLER
TUBES
IGNITOR
ATOMIZER
DOUBLE-ACTING
PISTON
OIL
PUMP
WATER
INLET
WATER
OUTLET
OUTPUT
SHAFT
SWASH
PLATE
GUIDE PISTOL
AND SLIDERS
ROLL SOCK
PISTON ROD
BURNER
ROTARY '
PREHEATER
HEATER
TUBES
Figure 5-7.
Philips Stirling Double-Acting Engine with
Swash-Plate Drive (Ref. 5-13)
-------
01
I
Figure 5-8. Parts Layout for Philips Stirling Double-Acting
Engine. Torino Application (Ref. 5-13)
-------
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00
Figure 5-9.
Front 3/4 View of Philips Stirling Double-Acting
Engine Mock-Up for a Ford Torino (Ref. 5-13)
-------
application to a Ford automobile. The design concept involves the use of
four pistons in cylinders linked in series by heater, regenerator, and cooler
units. Each piston functions alternately as a piston displacer and as a power
piston, leading to reduced weight and volume.
At the present time, the double-acting type of Stirling engine
is very much preferred over the single-acting, rhombic-drive Stirling engine.
The double-acting engines under development utilize two different types of
drive mechanisms. One drive mechanism is the common slider crank with
cylinders paired in a V formation. The other drive mechanism has cylinders
equally spaced around a circle with cross-head pistons acting on the surfaces
of a rotating wobble-plate or swash-plate mechanism. As Figure 5-10 shows
in schematic form, the cylinders are interconnected by sequential arrangement
of heater regenerator, and cooler. Figure 5-11 shows a schematic of the
four-cylinder, double-acting engine, illustrating the 90-degree phase relation-
ship between the cylinders. The cylinders could be connected to a V-crank,
as shown in Figure 5-12, or a swash-plate mechanism shown in Figure 5-13.
5.2.2 Design Features
5.2.2. 1 Power
As shown in Figure 5-14 the torque characteristics of Stirling
engines droop as speed is increased, somewhat like that of gasoline or diesel
engines. The Stirling engine torque droops naturally as speed is increased
due to reduced effective heat-transfer capacity and increased friction flow
losses of the working fluid. Torque drops off rapidly at lower speeds owing
to leakage past the power piston.
The maximum output speed of current Stirling engines tends
to be lower than that of conventional automotive gasoline engines. The double-
acting type engine being lighter in weight should accelerate to a given speed in
shorter time than engines with rhombic drive. Currently, throttle response
from zero load to full load is approximately 0.30 second.
5-14
-------
EXPANSION
SPACE
COLD
COMPRESSION
SPACE
COOLER
REGENERATOR
HEATER
Figure 5-10. Double-Acting Type Engine, Schematic (Ref. 5-3)
-------
(Jl
I
SINE WAVE
MOTION DEVICE
Figure 5-11. Double-Acting Piston (Ref. 5-1)
-------
AIR PREHEATER
HEATER
AIR
REGENER-
TOR
COOLER
PISTON
COMBUSTOR
ROD SEAL
CASING
Figure 5-12. Double-Acting V-4 Engine Preprototype (Ref. 5-4)
5-17
-------
oo
HEAT
SOURCE
COOLING
WATER
SWASH PLATE
Figure 5-13. Double-Acting Cylindrical
Configuration (Ref. 5-5)
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250 500 750 1000 1250 1500 1750
ENGINE SPEED rj, rpm
2000 2250 2500
Figure 5-14.
Dynamometer Torque of a 40-hp Stirling Engine
as a Function of the Speed r] at Different Values
of P max (Ref. 5-2)
-------
5.2.2.2 Specific Weight
On the basis of projected data involving development of the
Stirling engine for automotive purposes, the specific weight is estimated
as 3.5 to 4.5 pounds per hp (Ref. 5-12). This is based on an advanced-design
170-hp, swash-plate engine using hydrogen gas at a mean working pressure
of 3,000 psi with indirect heating and operating between 1,380°F and 131 °F.
This specific weight is for an engine that includes combustor and preheater
but excludes radiator, flywheel, and accessories .
5.2.2.3 Specific Volume
The double acting swash plate Stirling engine using hydrogen
gas shows promise of being fitted into the volume previously occupied by a
conventional Otto cycle engine in a 1972 Ford Torino (Ref. 5-5). A number
of changes are required in coupling the engine to the transmission (a 5-inch
rearward movement of the transmission for example) and in the suspension,
steering box, and dashboard (Ref. 5-5). A 170-hp, swash-plate engine for
this application is estimated to have a volume of approximately 10 ft 3 or a
specific volume of about 0.06 ft /hp. Recourse to the double-acting type of
Stirling engine has reduced the specific volume by a factor greater than two
compared to a rhombic-drive engine.
5.2.2.4 Materials
The Stirling engine uses most of the materials found in con-
ventional internal combustion engines. Cast iron can be used for the crank-
case, steel for the crankshaft, and aluminum for most non load carrying
members .
The costly material in the Stirling engine is in the heater
tubes. These tubes are made of heat resistant chrome and nickel alloys
such as SAE 30310 stainless or Multimet. The heater tubes are also costly
to fabricate. Further advances are being made in the design of heater tubes
to improve heat transfer.
5-20
-------
Several developers are testing the heat-pipe concept for use in
this engine. The concept makes use of an enclosed volume with a porous lining
filled with a heat-carrying-fluid medium such as sodium. At the hot end of the
heat pipe sodium is vaporized and allowed to flow to the cool end, where it
condenses (Ref. 5-6); it returns to the hot end by capillary action. The afore-
mentioned improvements permit the heat tubes to be shorter, thereby lower-
ing friction losses, increasing efficiency, and reducing the cost of materials
in the heater head. Also, the distribution of temperature is improved with the
use of sodium thereby lengthening the life of the heater head.
It must also be noted that the use of the double acting principle
permits one heater head to be used for a multicyUnder engine. This cuts
down the cost of fabricating heater heads, reduces heater head size, and
makes the engine more compact for automotive purposes.
5.2.2.5 Fuel Requirements
The Stirling engine is an external combuston engine and,
therefore, a multifuel engine. It can operate with any of the commercial
fuels sold on the open market and has no special octane or cetane require-
ments. Because the fuel is injected into the combustor, it is necessary that
the fuel be filtered to prevent clogging of the injector orifice. Commercial
fuels are highly filtered and, therefore, are preferred to noncommercial
fuels.
5.2.2.6 Temperatures and Pressures
The Stirling engine must be operated at high heater tempera-
tures and relatively low cooler temperatures if high efficiency is to be
achieved. Whereas the maximum cycle temperature of an Otto cycle engine
will reach 3,500 to4,000°F, the Stirling engine operates at a lower tempera-
ture. For example, the Ford Torino engine is designed for 1,470°Foperation
(Ref. 5-5). This is the practical operating temperature for heater tubes
fabricated from heat-resistant and creep-resistant stainless steels.
5-21
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5.2.2.7
Working Fluids
The preferred working fluid is hydrogen gas. This gas has
properties conducive to low friction losses and high heat transfer. Under
high working pressures and temperature, hydrogen can diffuse through the
metal container wall and thereby escape out into the atmosphere. A thin
coating of silicon nitride is used as a hydrogen gas diffusion barrier. It has
been claimed that the coating is effective under apparently static internal
pressurization conditions. It is not known whether the coating effectiveness
varies for materials undergoing deformation during creep and fatigue
conditions.
Helium gas is the next preferred fluid. It has low molecular
weight and does not diffuse through container walls. Helium is safe to handle
under both low and high pressure. However, thermal efficiency of the engine
is lower than that of the same engine using hydrogen gas as the working
medium (Figure 5-4). Air was used at one time, but the thermal efficiency
was lower than for either hydrogen or helium.
5.2.2.8 Cooling Requirements
Compared to an Otto or Diesel Cycle engine, the Stirling
engine transfers a much greater proportion of its heat through its coolant.
For this reason, the Stirling engine in certain designs may require a radiator
about 2. 5 times as large as that for an internal combustion engine of com-
parable power. The following table shows percentages of heat balances for
Stirling and diesel engines (Ref. 5-8).
Exhaust
Cooling Water
Friction
Auxiliaries
Effective
Diesel, %
35
19
7
2
37
100
Stirling, %
9
47
4
4
36
100
5-22
-------
EPA's research and development program in the Rankine
cycle power plant field has resulted in design and fabrication of a steam
condenser of smaller size than previously available for air cooled systems
(Ref. 5-9). This means that the radiator for a Stirling engine can be
reduced to a size compatible with automobile engine compartments by using
using the ideas involved in design of the aforementioned steam condenser.
It should be noted that the Philips Company has also developed, and is
further improving, a smaller size radiator for Stirling engine automobile
application (Ref. 5-3). This radiator is called a "Folded Frontal Area
Radiator" and consists of fine-gauge metal attached to small-diameter,
coolant-carrying tubes folded like an accordion to present a large frontal
area within a small volume.
The cost to manufacture the two types of aforementioned
radiators is likely to be higher than that for conventional-type radiators found
in present-day automobiles.
5.2.2.9 Transmission
The engine can use a conventional transmission, either
manual or automatic. A properly designed engine could provide high torque
at low speed with a relatively flat characteristic over the speed range. On
this basis, it might be anticipated that fewer gear changes would be required,
compared to an Otto cycle engine.
5.2.2.10 Lubrication
Because the Stirling engine is an external-combustion, low-
temperature engine, the lubrication oil does not operate in an environment as
hot as that found in the Otto cycle engine. Use of helium or hydrogen gas re-
quires special seals and/or use of a crosshead piston designed to restrict the
flow of oil into the hot-working-fluid area or flow of the working fluid into the
crankcase. The oil consumption has been found to be minimal in a Stirling
engine, and it would be advantageous to use the same oil in the transmission
as in the engine.
5-23
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5.2.2.11 Filtration
Filtration of combustion air is required to prevent fouling and
clogging of air passageways formed by the stacking of heater tubes. The
filter must be able to handle air at low pressure and at high preheat temper-
temperature.
Filtration of lubricating oil is required to prevent access of
abrasive materials to lubricated load carrying surfaces such as the bearings
and cylinder surfaces.
5.2.3 Operating Characteristics
5.2.3.1 Power Control Requirements and Methods
The output of a Stirling engine of given displacement is nearly
a linear function of the mean pressure of the working gas. Sudden changes
in mean pressure can effect sudden changes in engine output. This fact
forms the principle of control.
The Philips control system consists of two independent
operations: Fuel flow is controlled by the heat exchanger exit temperature,
which is affected by the pressure of the working fluid and heat input from the
combustor. Engine throttling is controlled by varying the pressure in the
chamber housing the working fluid.
A schematic of the pressure-regulating method is shown in
Figure 5-15 where a reservoir supplies high pressure gas to the working
fluid space. Throttle control is effected by opening and closing the supply
and short-circuit valves as required. With this system a compressor is
required to pressurize the gas and return it to the reservoir.
The power control system schematic is shown in Figure 5-16
and the fuel control system in Figure 5-17. With this method of control it
is possible to go from zero power to full power in 0.3 seconds.
5.2.3.2 Starting Characteristics
The Stirling engine has reliable starting qualities under all
conditions. Startup time from the cold condition is determined by design
5-24
-------
ro
Ul
RESERVOIR
SHORT-CIRCUIT
X , VALVE
SUPPLY VALVE
A
DUMP VALVE
COMPRESSOR
Figure 5-15. Schematic of Pressure-Throttling
Method (Ref. 5-3)
-------
cr-
COMPRESSOR
PI
HIGH PRESSURE
GAS RESERVOIR
ACCELERATOR
PEDAL
Figure 5-16. Power Control Diagram (Ref. 5-1)
-------
Ul
I
AIR INLET
VALVE
BURNER
COMBUSTION
AIR BLOWER
HEATER TUBES
TEMPERATURE
SENSOR
FUEL PUMP
FUEL/AIR CONTROL
Figure 5-17. Fuel Control Diagram (Ref. 5-1)
-------
variables. A startup period of 10 to 15 seconds can be achieved (Ref. 5-5);
however, necessary design compromises may require a period of up to 30
seconds. Starting characteristics are defined by the component parts of a
sequential control system which activates the combustion blower, fuel flow,
ignition, and engine cranking.
5.3 PERFORMANCE CHARACTERISTICS
5.3.1 Power and Efficiency
The application of the double-acting principle to the Stirling
engine has spurred development of four-cylinder engines with low values of
weight per horsepower. The specific weight of a Philips four-cylinder,
single-action, in-line, rhombic drive of 200 hp is 8. 8 pounds per horsepower.
The projected specific weight of the swash-plate,double-acting automotive
engine is about 3.5 to 4.5 pounds per brake horsepower.
The actual Stirling engine thermal efficiency at steady-state
maximum to minimum load ranges from about 28 to 48 percent (Figure 5-4).
The thermal efficiency decreases with increasing load, increasing rpm and
increasing molecular weight of the working fluid.
The three most used working gases are: air, helium, and
hydrogen. Use of hydrogen gas as a working fluid reduces internal-flow
friction and increases heat transfer rate. These gases are installed at
high mean pressures in the cylinders. The maximum mean pressure has
been continuously raised and is now at approximately 200 atmospheres
(Ford Torino engine). The performance map for the Ford Torino engine,
given in Figure 5-18, shows torque as a function of engine speed for several
pressure levels.
5.3.2 Emissions
The Stirling engine being an external combustion engine can
be expected to have exhaust emissions comparable to a Rankine cycle engine.
Very low mass emission levels of CO and HC have been reported for Stirling
engines. While NOx emissions are lower than for comparable uncontrolled-
5-28
-------
U1
I
ro
vO
TORQUE
kg-m
t
60
50
40
30
20
10
Pc = 200"atm
P = 175*atm
C I
P = 150 atm
" C I
P = 125 atm
\
PC= 100 atm
P_ = 75 atm
LP = 50 atm^-»
P^ = 25 atm ——? - —
C ^^
125 (kW)
100 (kW)
75 (kW)
50 (kW)
25 (kW)
10 (kW)
1000 2000 3000
—^ENGINE SPEED, rpm
4000
Figure 5-18. Performance Map of Ford Torino Stirling
Engines (Ref. 5-1)
-------
emission gasoline engines (as low as 1 percent of gasoline engine NOX
emissions), considerable variation exists between engines. Further reduc-
tion of NO emissions for some engines should be possible through directed
Jt
research. EPA's Alternative Engine Development Program has developed
a combustor for the Rankine cycle engine (Ref. 5-7) which has very low
exhaust emissions. Principles involved in the design of this combustor
are applicable to the design of a low exhaust emission combustor for the
Stirling engine.
Other data are available from test simulations by Philips.
The following table shows emissions for the double acting Stirling engine
(Ford Torino engine) with exhaust gas recirculation but without add-on
devices and excluding cold start conditions. The test results are presented
in Table 5-1 and compared with the 1976 Federal exhaust emission standards.
As indicated, all emission values are lower than the standard.
Table 5-1. CVS Test Simulation Emissions in gm/mile
(Ref. 5-10)
Gas
HC
CO
NO
X
N. V. Philips
0.10
0.31
0.175
Original 1976
Federal Standard
0.41
3.40
0.4
5.3.3
Fuel Economy
The fuel economy of the Stirling engine is better than that of
the conventional high-compression-ratio internal combustion engine and is
comparable to that of diesel engines. Computed gasoline fuel economy for
the 175-hp swash-plate Stirling engine is compared with a 1976 Ford 351, V-8
engine in a same-weight vehicle, Table 5-2. It should be noted that the
5-30
-------
overall efficiency, and thus fuel economy, of the Stirling engine is particu-
larly dependent upon the coolant temperature. The calculated fuel economy
given in this table assumed a coolant temperature of 163°F. A higher tem-
perature resulting from an inadequate radiator or high ambient temperature
would degrade the Stirling engine efficiency and, hence, its fuel economy.
Table 5-2. Fuel Economy Comparisons (Ref. 5-1)
Consumer Average
Suburban Driving
City Traffic
Superhighway
1976 Ford V-8
10.7 mp g
13.3 mp g
8.2 mp g
13.0 mpg
Stir ling -Torino
(Projected)
14.7 mpg
17.9 mpg
11.4 mpg
16.6 mpg
Gain
37%
35%
39%
28%
5.3.4
Noise Levels
The Stirling engine noise level is about 10 to 15 dB(A) lower
than for current passenger cars and about 15 to 20 dB(A) less than a normally
silenced diesel engine. The noise level of a Philips Stirling engine installed
in a bus was measured to be 68 dB(A). The largest noise sources are the
combustion air blower and radiator fan. The absence of poppet valves,
intermittent combustion noise, and high cylinder-pressure gradients contrib-
ute to its quiet operation. The noise level of a double-displacement Stirling
engine is lower than that of the rhombic-drive engine because of the absence
of synchronizing gears, which are the major noise source in a displacer
engine.
5.3.5 Odor and Particulates
The Stirling engine exhaust odor is undetectable at concentra-
tions normally encountered in the atmosphere. Smoke intensity is almost
zero when the air-to-fuel ratio is greater than 26 to 1. Application of
exhaust gas recirculation (EGR) reduces the NOX emission and the CO and
5-31
-------
soot emissions at low loads . Increasing EGR at low loads does not detract
powerwise, since the blower consumes little power at this condition.
5.3.6 Maintainability and Reliability
The Stirling engine's heat tubes, anti-leak seals and control
system are the critical items with regard to maintainability and reliability.
(A unique roll-sock seal has been developed by Philips to separate the
working fluid from the lubrication system. Durability statistics are not
completed for this device at this time.) These items can be investigated
most meaningfully by operating a Stirling engine in a vehicle over the
Federal Driving Cycle for at least 50, 000 miles. At present, only laboratory
data are available and do not represent environmental conditions as they
are found in road operations.
5.3.7 Safety
A vehicle with a Stirling engine should be about as safe to
operate as a vehicle equipped with a diesel engine. Both engines inject
fuel into the combustion chamber, although the Stirling engine utilizes lower
injection pressures. If the working fluid is hydrogen gas, some danger
exists if the fluid escapes and collects in an enclosure where an explosion
could occur. However, hydrogen gas diffuses rather rapidly; moreover,
only a small quantity of gas is contained in the engine. If helium gas were
used as the working fluid, the Stirling engine would be safer to operate than
a diesel engine.
Toxicity is not present in the fluids currently considered for
use in a Stirling engine; hence, it is not considered a safety problem.
5.3.8 Driveability
The Stirling engine has torque and speed characteristics
comparable to current spark ignition engines (Ref. 5-5). Therefore, drive-
ability of the vehicle should be similar to that of a vehicle powered by the
5-32
-------
spark ignition engine. Table 5-3 shows performance projections for the
Ford Torino powered by a Stirling engine and another powered by a spark
ignition engine. As shown, performance of the two engines is comparable.
Table 5-3. Performance Projections (Ref. 5-5)
Distance, 0-10 Sec. ft
Time, 0-60 Mph, sec
Time, 25-60 Mph, sec
Time 50-80 Mph, sec
Stirling 4-215
Engine
491.0
11.1
8.0
11.2
354-2VC Passenger-
Car Engine
484.0
11.1
8.0
11.3
5.4
5.4.1
CURRENT STATUS OF TECHNOLOGY
Current Utilization
The major research work on the Stirling engine was
accomplished by Philips Research Laboratories of Holland. Philips has
established that a double-acting engine with a swash plate is suitable for
installation in a contemporary medium-size automobile.
A combined development program conducted by Ford and
Philips has resulted in the following technology status for the Stirling engine
(Ref. 5-1):
a. The Stirling engine has potential for low-emissions operation
in a current domestic automobile.
b. The engine has the potential for very low noise level of
operation.
c. The Stirling engine performance is similar to that of a
spark ignition engine.
d. The Stirling engine has the potential for excellent fuel
economy.
e. The Stirling engine can be packaged within the present-day
engine compartment.
5-33
-------
Ford, in addition to work with Philips, has an agreement with United Stirling
of Sweden to develop a small V-4 Stirling engine for powering a Pinto car.
5.4.2 Current Research and Development
Current research and development activity is directed at
proving the potential of the Stirling engine as an automobile power plant.
Major open issues in Stirling engine development are:
a. Cooling system requirements
b. Use of hydrogen as a working fluid and for seal durability
c. Maintenance to ensure low emissions for 50, 000 miles
d. Mass-production manufacturing cost.
Ford and Philips will concentrate on solving these problems and expect to
install a prototype engine in a 1975 Ford Torino for performance verification
road tests (Ref. 5-1).
5.5 PROJECTED STATUS
5.5.1 Potential for National or Regional Transportation
The Stirling engine was developed initially for special appli-
cations where performance was of utmost importance and manufacturing
cost considerations were secondary. As stated in Ref. 5-11, the Stirling
engine could, technically, replace the spark ignition engine; however, mass
production of this engine will be delayed until economic feasibility can be
proven. The engine cannot compete with the cost of the present day spark
ignition engine.
Two scenarios could lead to increased emphasis for this
engine as a candidate for future automobiles:
a. The cost of fuel increases to a point where the improved
fuel economy of the Stirling engine offsets the higher first
cost and its lifetime cost results in an overall economic
advantage over the spark ignition engine.
b. Air pollution legislation dictates the use of a low emission
engine like the Stirling engine.
5-34
-------
In any event, even at an accelerated rate of research, it will be mid-1980
before this engine can become available for limited production.
5.5.2 Impact on Energy Requirements and Air Quality
As previously discussed, this engine is highly efficient
and provides low emission levels. Therefore, it is an ideal engine from
the viewpoint of combining conservation of energy with improvement in
air quality.
5-35
-------
SECTION 6
-------
6. DIESEL ENGINES
6.1 INTRODUCTION
6.1.1 General Description
The diesel or compression ignition engine is a reciprocating
engine which is, in many respects very similar to the Otto cycle spark ignition
engine. In the diesel, air alone is compressed in the cylinder, and the fuel is
then injected into the heated air towards the end of the compression stroke.
The temperature developed during compression is sufficiently high to ignite
the fuel immediately after injection. Thus, spark plugs, distributor, and car-
buretor are eliminated in the diesel, but a high-pressure, fuel-injection sys-
tem is required. Since compression-ignition demands high pressure ratios,
the current diesel engines are heavier, more bulky, and costlier than equiva-
lent spark ignition engines.
The principal advantages of diesels relative to spark ignition
engines include their ruggedness and their higher thermal efficiency, which re-
sults in lower fuel consumption, particularly at part-load operating conditions.
Many different engine configurations have been built around
.the basic thermodynamic diesel cycle, and some of these are currently
being marketed by many companies throughout the world. The various
designs can be grouped into two categories -- open chamber and divided
chamber -- and each of these can then be subdivided into different engine
classes, depending upon specific features related to the design of the com-
bustion chamber and the induction system. Open-chamber designs are favored
by most manufacturers of heavy-duty diesels because of the better fuel
economy obtained with this engine type. Conversely, divided-chamber
designs (prechambers and swirl chambers) are preferred by the manufactur-
ers of light-duty engines because of their higher speed capability.
6-1
-------
6.1.2 Historical Background
The dies el engine was invented by Rudolf Diesel in 1892 and was
originally intended to burn coal dust. In his early designs, Diesel proposed
to coordinate the rate of fuel injection with the movement of the pistons in
such a manner that the heat of combustion would be liberated at constant
temperature, thus approaching the thermal efficiency of the Carnot cycle.
However, this approach proved to be unsuccessful, and the cycle was then
modified to achieve near-constant pressure combustion. While this ideal
process is still being utilized in large, low-speed diesels, the high-speed
dies els used in light-duty applications employ a process between the ideal
cycle and the Otto cycle.
The high compression ratio required to achieve fuel ignition,
and the resulting bulk, precluded the application of diesels in early road
vehicles. The first dies el-powered truck was built in 1907 by Saurer in
Switzerland, in cooperation with Diesel; in 1911 the first fully loaded
dies el truck crossed the United States from coast to coast.
Daimler-Benz of Germany started series production of its
Mercedes Benz 260D diesel-powered automobile in 1936. Since resumption
of production in 1949, Daimler-Benz has sold more than one million diesel
automobiles. A number of other companies have since entered the light-duty
diesel engine field, including Peugeot of France, Perkins of England,
Daihatsu of Japan, and Opel of Germany.
6.2 POWER PLANT DESCRIPTION
6.2.1 Diesel Cycle
The thermodynamic process describing the operation of the
ideal diesel, or compression-ignition engine, is depicted in Figure 6-1 in
terms of the cylinder pressure versus cylinder volume and the temperature
of the combustion products versus system entropy. In the ideal cycle, the
compression and expansion processes are isentropic processes; combustion
is a constant-pressure process, and the blowdown of the exhaust gases is
a constant-volume process.
6-2
-------
In the ideal dies el cycle, air is inducted into the cylinder
during the intake stroke of the piston and compressed isentropically during
process (jf) - (2) (Figure 6-1). Fuel injection into hot compressed air and ig-
nition occur during process (I) - (|) . Injection and ignition take place in such
a manner that the descending piston compensates for the pressure buildup
that would otherwise occur as a result of the heat released during combustion.
The combustion process is completed at point QQ an(i the high pressure gases
then expand isentropically to point Q£), thereby performing additional work on
the piston. The exhaust valve opens at point @ , and the bulk of the exhaust
gases discharge at nearly constant volume. The remaining exhaust gases
are then pushed out by the ascending piston.
The indicated thermal efficiency of the ideal dies el cycle is
presented in Figure 6-2 for two specific heat ratios, k = 1.3 and 1.4. Also
shown for comparison are curves representing the ideal Otto cycle. For a
given compression ratio and specific heat ratio, the indicated thermal effi-
ciency of the dies el cycle is lower than that of the Otto cycle. However,
diesels are normally designed to operate at much higher compression ratios
(of the order of 15:1 to 20:1) than spark ignition engines (typically 9:1), and
this improves the efficiency of the diesel with respect to the spark ignition
engine. Also, the specific heat ratio of the diesel combustion products is
higher than for spark ignition engines, especially under part-load conditions
which further improves the diesel cycle efficiency. These effects are respon-
sible for the higher efficiency of diesel engines relative to gasoline engines.
In actual diesels, the combustion process is modified
somewhat from the ideal process illustrated in Figure 6-1. In the real
engine, a substantial portion of the total fuel is injected slightly before top
dead center (TDC), and combustion of that fuel occurs at essentially constant
volume, with the piston at TDC. Combustion of the remainder of the fuel
takes place at nearly constant pressure while the piston starts to descend
from TDC. The indicated thermal efficiency of this "dual cycle" process
falls between the diesel cycle and the Otto cycle.
6-3
-------
Fuel Injection & Combustion
Adiabatic
Compressi
(air only)
min-
"mm
Exhaust
max
Volume
(U
a
Fuel Injection &
Combustion
Ignition
GJ
r
Isobar
Adiabatic
Compression
(air only)
Isochore
Adiabatic Expansion
Smax Entropy
Figure 6-1. Ideal Diesel Cycle
6-4
-------
O.SO
0.70
o
z
UJ
u.
UJ
-I
OTTO CYCLE
^SPECIFIC HEAT
RATIO k = 1.4
0.50
ex
UJ
I
I-
0.40
0.30
DIESEL CYCLE
k= 1.4
k= 1.3
10
15
COMPRESSION RATIO
20
25
Figure 6-2 .
Diesel and Otto Cycle Thermal Efficiency
Compres sion Ratio
vs
-------
6.2.2 Power Plant Configuration
Although all diesel engines are based on the thermodynamic
cycle illustrated in Figure 6-1, many different engine configurations have
been designed around this cycle, and a number of these are currently being
marketed by many manufacturers. The various engine designs can be
grouped into two categories: (1) open-chamber or direct-injection engines,
and (2) divided chamber or indirect-injection engines. Each of these two
engine categories can then be subdivided into a number of engine classes,
depending upon the type of combustion chamber and air induction system
utilized.
Open-chamber engines have a single combustion chamber per
cylinder, and the fuel is injected into this chamber and vaporized for subse-
quent combustion (Refs. 6-1 through 6-3). The following three engine classes
fall into this category:
a. Quiescent or low-swirl "Mexican hat" chambers, favored
by many domestic manufacturers (Figure 6-3)
b. Medium-swirl, deep-bowl chambers, used primarily
in Europe (Figure 6-4)
c. High-swirl, spherical "M" combustion chambers used
in special applications (Figure 6-5).
The divided-chamber engine category employs a main com-
bustion chamber, formed by the cylinder walls, the piston and the cylinder
head, and an antechamber which is connected to the main chamber through
a communicating flow passage. This engine category is further divided into
three classes, as follows:
a. Swirl chamber or turbulence chamber, a configuration
favored by a number of manufacturers of diesel passenger
cars, including Peugeot and Opel (Figure 6-6)
b. Precombustion chamber, or prechamber. This design
is used in the Mercedes Benz 220D diesel-powered
automomobile (Figure 6-7)
c. Air cell and energy-cell combustion chamber (Figure 6-8)
6-6
-------
Figure 6-3. Low-Swirl Diesel Engine (Ref. 6-4)
6-7
-------
Figure 6-4. Medium-Swirl Diesel Engine (Ref. 6-5)
FUEL SPRAY
INT. X O A—. AIR MOTION
EXH.
Figure 6-5. High-Swirl, "M" System Diesel Engine
(Ref. 6-6)
6-8
-------
NOZZLE
PLUG
(optional)
Figure 6-6. Ricardo Swirl Chamber Diesel (Ref. 6-7)
Figure 6-7. Daimler-Benz Precombustion Chamber Design
(Ref. 6-8)
6-9
-------
AIR CELL
NOZZLE
Figure 6-8. Air-Cell Diesel Engine Schematic
(Ref. 6-1)
Because of lower fuel consumption (approximately 5 to 8
percent), direct injection engines are favored by many manufacturers for
use in industrial and truck applications. However, the divided-chamber con-
figurations are preferred by the makers of automotive diesels and by some
of the truck engine manufacturers because of the higher speed capability and
the lower emission potential of this particular engine design.
Compared with open-chamber diesels, the divided-chamber
engine has a number of additional inherent advantages, as well as disadvan-
tages. Because of its higher volumetric efficiency, the engine can be oper-
rated at higher brake mean effective pressures. This improvement is re-
alized by employing near-zero intake swirl. Furthermore, at a given
compression ratio, the maximum pressures, temperatures, and pressure-.
rise rates in divided-chamber engines are significantly lower than for open-
chamber designs. As a result, the engine noise, exhaust emissions, and
mechanical stress characteristics of the engine are improved accordingly.
6-10
-------
The principal disadvantages of divided-chamber diesels
compared with open-chamber configurations include higher fuel consumption
and higher exhaust gas temperatures. The loss in fuel economy is primarily
due to the higher surface-to-volume ratio of the engine and, in the case of
the swirl chamber configuration, the high swirl velocities in the combustion
chamber.
Daimler-Benz has chosen the low-swirl prechamber design for
its 220D passenger car for a number of reasons, including the high-speed
capability of the engine and its high volumetric efficiency and low noise and
emission levels. To enhance ignition at idle and low loads, Mercedes utilizes
a ba.ll pin located in the center of the prechamber. Upon ignition, this pin
heats up very rapidly and maintains a temperature of about 1300"F under all
operating conditions. The fuel injected into the chamber impinges on the
ball pin, thereby providing a stable ignition source and assuring brief igni-
tion delays at all times. During cold start of the engine at ambient tempera-
tures below about 70 F, activation of the electrically heated glow plug is
required for a few seconds in order to minimize engine cranking.
6.2.3 Design Features
6.2.3.1 Power Output
The specific power output of heavy-duty and light-duty diesels
in terms of maximum horsepower per cubic inch displacement is presented
in Figure 6-9 as a function of maximum horsepower. Also shown are corre-
lations established for domestic and imported 1971 model year light-duty
gasoline engines (Ref. 6-9). The specific power output of the smaller spark
ignition engine used in imported cars is about 40 to 50 percent higher than
for domestic passenger car engines. This trend is the direct result of the
higher maximum speed of the imported engines. As indicated in the figure,
the specific power output of the passenger car diesel engines manufactured
by Daimler-Benz and Opel is only slightly lower than that of 1971 domestic
6-11
-------
1.2
Q 1.0
u
a.
K 0.8
D
GL
°0.6
UJ
o
a
O 0-4
O
LU
0.2
a
1971 IMPORTED PASSENGER CAR
GASOLINE ENGINES
O O
OOPEL 2100D DIESEL
A MERCEDES 220D DIESEL
A
1971 DOMESTIC PASSENGER
CAR GASOLINE ENGINES
TRUCK DIESELS
(4 STROKE)
100 200
MAXIMUM POWER OUTPUT, hp
300
400
Figure 6-9.
Specific Power Output of Diesel and Gasoline Engines
(Maximum Horsepower/Cubic Inch Displacement)
-------
spark ignition automobile engines, but only about 60 to 70 percent of the
specific power of the smaller, higher-speed engines used in imported
vehicles.
A number of factors contribute to the relatively low specific
peak-power output of current diesels, in spite of the higher thermal efficiency
of the diesel cycle. These include (1) the lower maximum speed capability
of the diesel which is limited by the stress loading of the pistons, connecting
rod, crankshaft, and bearings, (2) the lean air-fuel ratio operation, and
(3) the loss in volumetric efficiency occurring at high engine speeds
(Ref. 6-12). In divided-chamber diesels, additional power losses are due
to the pressure drop occurring in the flow passage connecting the two
chambers. However, these losses are partially compensated for by the
higher compression ratio generally used in divided-chamber engines.
The power output of diesel engines can be substantially
increased by means of turbocharging or supercharging. As a result, the
specific power output of the engine increases accordingly and can exceed
that of emission-controlled gasoline engines. Turbochargers are frequently
utilized in truck applications to increase the peak power output of the engine,
especially at altitude, and to improve its fuel economy. However, in the low
power regime of the engine, the turbocharger is generally much less effec-
tive, and its response to load variations of the type encountered in passenger
cars is rather slow because of the inertia of the turbine and compressor
rotors. These problems are alleviated to a large extent in the Comprex
supercharger system, which has been developed by Brown Boveri of Switzer-
land for use in heavy-duty diesel engine applications (Refs. 6-10 and 6-11).
This device, shown schematically in Figure 6-10, is a compression wave
machine which utilizes the energy of the exhaust gases to compress the
intake air before admission into the cylinder. Relative to turbochargers,
the Comprex system has a number of advantages, including (1) fast response
to engine load variations; (2) higher BMEP capability at low engine speeds;
(3) "internal" EGR, which can be achieved by proper timing of the intake
valve closing; and (4) lower smoke and NO emissions. The disadvantages
.X
6.13
-------
EXHAUST
AIR INTAKE
Figure 6-10. Comprex Supercharger Arrangement
(Ref. 6-10)
6-14
-------
of the Comprex system include its larger size and higher weight, cost, and
noise levels.
6.2.3.2 Specific Weight
Specific weight data (pounds per horsepower output) for current
automotive and industrial open-chamber and prechamber diesels are plotted
in Figure 6-11. The open-chamber data were utilized to establish the trend
of specific weight with rated horsepower. The correlation shown for divided-
chamber engines was then obtained by transferring that slope to the mean
level of the divided-chamber data. The two curves show that divided chamber
engines are of the order of 25 percent lighter than equivalent open-chamber
designs. This weight difference is directly related to the lower peak combus-
tion pressures developed in divided chambers resulting in weight savings in
the cylinder head, engine block, connecting rods, crankshaft, and bearings.
Typically, the peak pressures in divided-chamber engines are between about
1,000 psi and 1, 200 psi, compared with about 1,400 psi to 2, 000 psi for open-
chamber configurations. The peak combustion pressures in spark ignition
engines are of the order of 600 psi for low compression ratio engines, and
as high as 1,000 psi for high-compression-ratio designs.
The specific weight of automotive reciprocating spark ignition
engines is also presented in Figure 6-11. This correlation is based on
25 engines ranging from 50 to 350 brake-horsepower output. The weights
shown include the starter, alternator, and flywheel, but exclude the radiator
and the cooling water (Ref. 6-14).
Comparison of the data indicates that current diesels are con-
siderably heavier than equivalent gasoline engines. For example, a 100-
horsepower, naturally-as pirated, divided-chamber diesel weighs about
620 pounds compared with only 350 pounds for the gasoline engine. With
40 percent turbocharging, the weight of this engine can be reduced to
about 500 pounds.
Additional weight savings might be possible by (1) increasing
the degree of turbocharging, (2) reducing the peak combustion pressures,
and (3) lowering the design life of the diesel engine. Currently, the life of
6-15
-------
50
a
JC
O
uj 10
O
UJ
a.
DIVIDED CHAMBER
DIRECT INJECTION
AUTO.
A
•
IND.
A
D
A OPEN CHAMBER
B DIVIDED CHAMBER (DC)
C DC TURBOCHARGED
D SPARK IGNITION
10
RATED ENGINE HORSEPOWER
Figure 6-11. Specific Weight of Diesel and Spark Ignition Engines
(Ref. 6-13)
-------
truck diesels before major overhaul is of the order of 300, 000 miles or
more, compared with about 80, 000 miles for light-duty gasoline engines.
The specific weight of newly designed light-duty diesels is estimated to be
comparable to that of gasoline engines (Ref. 6-15).
6.2.3.3 Specific Volume
3
The specific volume (ft /hp) of current heavy-duty dies el
engines is shown in Figure 6-12 as a function of rated engine power. There is
no reason to expect biases between open and divided chambers and no attempt
was made to correlate the data on that basis. The turbocharged engine corre-
lation shown (Figure 6-12 was drawn on the basis of a 40 percent realized
power increase, compared with naturally aspirated diesels. Also presented
in Figure 6-12 is the specific volume curve for reciprocating spark ignition
engines. Both curves are based on the total engine envelope from fan to fly-
wheel and from air cleaner to oil pan.
As indicated Figure 6-12, current diesel engines are consider-
ably larger than equivalent gasoline engines. For example, a 100-horsepower,
3
naturally-aspirated diesel engine has a volume of about 17 ft compared with
about 10 ft for the gasoline engine. However, the size of the diesel could be
reduced substantially by turbocharging or supercharging.
6.2.3.4 Manufacturing and Materials
Basically, the fabrication techniques, tolerances, tooling, and
materials used in the manufacture of automotive diesels are quite similar to
those employed in gasoline engines. However, according to Daimler-Benz,
some modifications and additions to the transfer lines of spark ignition engines
are needed to account for the additional steps required in the manufacture of
diesel engines (Ref. 6-12). Although the gasoline and diesel engines
utilized by Mercedes in its 220 and 220D vehicles were originally designed
to use common parts whenever possible, there are many specific design
differences. These include the cylinder head (which is cast iron in current
diesels and aluminum in the spark ignition engine), engine block, head gasket,
pistons, rings, crankshaft, intake manifold, and starter. Similar materials
6-17
-------
10
I
t—«*
00
Q.
.C
LU
lio-1
o
o
u.
u
UJ
Q.
10
-2
O
NATURALLY ASPIRATED DIESEL
TURBOCHARGED
DIESEL
SPARK IGNITION
• AUTOMOTIVE
O INDUSTRIAL
10
MAXIMUM ENGINE HORSEPOWER
10"
Figure 6-12. Specific Volume of Diesel and Spark Ignition Engines
(Ref. 6-13)
-------
and fabrication procedures are currently utilized by the other manufacturers
of light-duty and heavy-duty diesel engines.
6.2.3.5 Fuel Requirements
Although most diesels run quite well on a variety of fuels, cer-
tain fuel properties are required to achieve optimum conditions in terms of
engine performance and smoke emissions. Cetane number is by far the most
important fuel parameter. For good combustion, the cetane number, which
is a measure of the autoignition characteristics of the fuel, should be suffi-
ciently high to minimize the ignition delay time. Cetane numbers and
autoignition temperatures of a number of fuels are listed in Table 6-1.
Experience has indicated that little improvement in the combustion process is
obtained for cetane numbers above about 50 or 55 (Ref. 6-17).
As indicated in Table 6-1, the autoignition temperature of
low-cetane fuels is high, making ignition more difficult to achieve. There-
fore, diesel engines designed to operate with these fuels generally have a
higher compression ratio to ensure ignition and stable combustion during
starting and at low power conditions. Since the occurrence of ignition in
diesels depends on the heat transfer into the fuel droplets, the maximum
feasible engine speed decreases with decreasing cetane number. Preheating
the air charge counteracts that problem to a large degree.
The sulfur in the fuel has a significant effect on the life of the
cylinder liner, piston rings, and bearings of reciprocating engines. When
using high-sulfur fuels, frequent oil changes are desirable to counteract the
problems related to acid formation. For example, if the sulfur content
increases about 0.5 percent, Daimler-Benz reduces the oil change interval
for its light-duty automotive diesel engine from 3,000 miles to 1,500 miles.
With respect to fuel effects on diesel exhaust emissions, a
considerable amount of work has been conducted by several investigators.
Data published by Marshall and Hurn (Ref. 6-18) indicate very little change
in the hydrocarbon, carbon monoxide, and oxides of nitrogen emissions of
several diesel engines when operated on No. 1 and No. 2 diesel fuel and on
two different solvents. Similar results were obtained by Daimler-Benz for
6-19
-------
Table 6-1. Cetane Numbers and Autoignition
Temperatures of Different Fuels (Ref. 6-16)
Fuel
Ether
Cetane (C1^H34^
n-Heptane (C-,H,/)
Diesel Fuel A
Diesel Fuel B
Lubricating Oil,
30 SAE
Kerosene
Gasoline
Low Octane
High Octane
Methanol
Ethanol
Benzene
a-Methylnaphthalene
(CUH1())
Cetane
Number
100
100
60
55
45
45
37
20
10
10
10
3
0
ti(F
Ignition
Temp, at
450 psi
330
380
380
450
480
460
550
570
750
720
640
840
850
diesel fuels refined both in Europe and in the United States (Ref. 6-12). Con-
versely, tests conducted by Shahed, et al. (Ref. 6-19) on a four-stroke,
single-cylinder engine (4. 75-inch bore, 4. 75-inch stroke, compression
ratio 14. 2 : 1) using three different fuels with cetane ratings of 40, 52, and
56 indicate some reduction in NOx and specific fuel consumption as cetane
number is increased.
6-20
-------
6.2.3.6 Fuel Injection System
In diesel engines, the fuel injection system assumes the
function of the carburetor used in gasoline engines. It is probably the most
important component of the engine with respect to combustion performance,
emissions, and.noise generation, especially in large, open-chamber engines .
With increasing swirl rates and air turbulence, the effects of the fuel injection
system characteristics on combustion become less pronounced.
The fuel injection system consists of a high-pressure fuel pump,
a fuel distributor, and a series of injection nozzles. Typical injection pres-
sures are in the range of 2, 000 psi to 30, 000 psi. The system is designed to
provide a high degree of fuel atomization and distribution within a sharply
defined injection period. As further discussed in Section 6.3.3, injection
timing has a very strong effect on NO and smoke emissions.
In general, fuel injection systems have the capability to supply
a precisely metered amount of highly atomized fuel to the engine cylinder at
all times. Obviously, this has a very beneficial effect on fuel economy as
well as on HC and CO emissions.
6.2.3.7 Cooling Requirements
The cooling system utilized in diesel engines is very similar
to that of gasoline engines and normally consists of a radiator, a belt-driven
cooling fan and water pump, and a the'rmostatic control device.
In divided-chamber diesels, the surface-to-volume ratio of the
two combustion chambers is higher than in open-chamber engines, resulting
in high thermal loading of the cylinder head and pistons. Although design
solutions to the thermal loading problems are available for the cylinder head,
the overall heat flux into the engine block must be handled by the cooling sys-
tem (Ref. 6-20). In general, the radiator size required for divided chamber
diesels is about 30 percent to 50 percent larger than for equivalent open-
chamber diesels.
6-21
-------
6.2.3.8 Lubrication System
The lubrication systems used in automotive diesel and gasoline
engines are similar, except for an additional oil circuit required in diesels for
lubrication of the fuel injection pump. Typically, the oil flows from the sump
through a strainer and filter and then into the main passages of the engine
block.
Diesel engines require more stringent lubricating oil speci-
fications than gasoline engines, especially with respect to its detergent
properties, and acid neutralization and antiwear characteristics. Also,
adherence to the oil viscosity grade recommended by the manufacturer is
more important in diesels than in gasoline engines because of a more rapid
increase in the oil viscosity with mileage accumulation. This increase is
directly related to the higher blowby normally obtained in diesels. There-
fore, oil changes are recommended more frequently for diesels than for
gasoline engines. For instance, the recommended intervals for oil and oil
filter changes are 3,000 miles for the diesel-powered Mercedes 220D and Opel
Rekord 2100D automobiles. Peugeot recommends an even shorter oil change
interval of 1,500 miles.
6.2.3.9 Transmissions
Both manual and automatic transmissions are feasible for use
in diesel-powered automobiles. The Daimler-Benz 220D and the Opel Rekord
2100D passenger cars are available with either transmission type, while the
Peugeot 504 diesel is offered with a four-speed manual transmission only.
The transmissions of these vehicles are similar in design and construction
to those used in the corresponding gasoline-powered cars marketed by these
companies.
The automatic transmission by Daimler-Benz consists of a four-
speed planetary gear set and a hydraulic clutch. Gear selection is controlled
by engine speed and accelerator pedal position. The gears of this transmission
have been optimized with respect to engine torque and speed.
6-22
-------
Opel's automatic transmission represents a modification of
the basic unit used in its spark ignition engine powered Rekord models. The
design changes include the torque converter casing, oil pump, and band servo
spring. In the absence of an intake throttle in the Opel diesel, the vacuum
required for the gearbox control system is provided by a separate vacuum
pump.
6.2.4 Operating Characteristics
6.2.4.1 Power Control Requirements and Methods
With few exceptions, heavy-duty diesel engines operate without
intake throttle, while light-duty engines generally utilize intake throttling.
In the absence of an air-flow control valve, the power output of the diesel is
controlled entirely by the amount of fuel injected into the cylinders. At idle,
the engine operates at air fuel mixture ratios of the order of 150:1, and the
idle fuel consumption is only about 25 percent to 40 percent of that of gasoline
engines. With increasing power demand, the fuel flow rate is increased by
moving the accelerator pedal which is directly connected to the fuel injection
pump. At full power, the air fuel ratio of automotive diesels varies between
20:1 and 25:1, depending upon the smoke characteristics of the particular
engine and the permissible smoke levels.
6.2.4.2 Starting Characteristics
In general, the starting characteristics of diesel engines, when
warm, are quite comparable to spark ignition engines. However, difficulties
can arise during cold starts, particularly at very low ambient temperatures.
In divided-chamber diesels, the starting problem can be alleviated by using
an electronically heated glow plug in each prechamber. The purpose of the
glow plug is to heat the air in the prechamber and to provide a hot spot for
initial ignition. According to Daimler-Benz (Ref. 6-12), the degree of
"preglowing" needed before engaging the starter motor varies with ambient
temperature and, more importantly, with the temperature of the engine.
Typically, with a cold engine and an air temperature of 68 F, the required
6-23
-------
preglowing time is between 5 and 10 seconds. However, under extremely
low engine and ambient temperatures, as much as one minute of preglowing
might be necessary. Unlike gasoline engines, diesels are less likely to stall
after a successful start.
To provide sufficient energy for starting, the battery used
in the Mercedes 220D automobile is about 30 percent larger than.that of the
gasoline engine powered 220 vehicle. Also, the starter motor in the 220D
is more powerful (2.4 hp vs 1.5 hp for the gasoline engine) to compensate
for the higher compression ratio of the diesel.
The starting time of open-chamber and divided-chamber
diesel engines can be reduced by incorporating certain starting aids, at the
expense of added complexity and cost. These devices include intake flame
heaters, ether injection, and preinjection into the cylinders or intake mani-
fold (Ref 6-1). Other approaches that have been considered include removal
of the air swirl during engine starting (Ref. 6-6) and incorporation of variable
compression ratio pistons (Ref. 6-14).
6.3 PERFORMANCE CHARACTERISTICS
6.3.1 Power, Speed, Torque, Efficiency, Combustion
Representative performance curves for a light-duty automotive
diesel engine, operated at maximum fuel flow rates, are depicted in
Figure 6-13. The power, torque, and specific fuel consumption (SFC) curves
shown were computed from data published by Daimler-Benz for its 2-liter
displacement (121.3 CID) prechamber engine which was used in the Mercedes
200D automobile (Ref. 6-8). For comparison, wide-open throttle power and
torque curves for the Mercedes 2-liter spark ignition engine (not modified
for emission control) are also presented in Figure 6-13.
The curves clearly illustrate the lower power output capability
of the nonsupercharged diesel engine. In the low speed regime, the torque of
the Mercedes 200D diesel is only about 20 percent lower than that of the un-
controlled Mercedes 200 gasoline engine. This explains the rather good low-
speed acceleration capability of the 200D vehicle compared with equivalent
6-24
-------
Q.
.a
120.—
100
80
60
ui
z
O 40
Z
UJ
20
TORQUE
HORSEPOWER
100
80
60
UJ
•D
O
g
s_
.c
0.60
0.50
O
0 1000 2000 3000 4000 5000 6000
ENGINE, rpm
NOTE: 2 LITER DIESEL OM 621
2 LITER SPARK IGNITION ENGINE (uncontrolled)
Figure 6-13. Diesel and Gasoline Engine Performance Parameters,
Daimler-Benz Automobile Engines (Ref. 6-8)
6-25
-------
gasoline engine powered cars. However, the torque decrement increases
with increasing engine speed. For example, at the maximum speed of the
2-liter diesel (4,350 rpm), the torque of this engine is only slightly above
50 percent of that of the gasoline engine (Ref. 6-8). Of course, the torque
of the diesel engine could be substantially increased by incorporating a tur-
bocharger or supercharger.
The specific fuel consumption (SFC) of the Mercedes 2-liter
diesel and uncontrolled gasoline engines is presented in Figure 6-14 as a
function of engine speed and brake mean-effective pressure (BMEP). As in-
dicated in the figure, minimum SFC of the diesel is only slightly lower than
that of the gasoline engine. However, at low engine loads, the diesel exhibits
superior fuel consumption characteristics. This is the direct result of the
very lean air-fuel mixtures at part load, the lower pumping losses, and the
higher compression ratio of the diesel. As discussed in Section 6.2. 1, each
of these factors has a beneficial effect on the thermal efficiency of the diesel
cycle.
6.3.2 Diesel Combustion
The diesel combustion process is preceded by an ignition delay
period, which occurs between the injection of the fuel into the compressed
hot-air charge and the initiation of combustion (Refs. 6-21 and 6-22). This
delay time, composed of physical delay and chemical delay, is dependent
upon the fuel, the charge temperature, and certain engine design
characteristics.
In prechamber and swirl chamber diesels, fuel injection is
timed so that combustion commences near TDC for optimum performance.
Under these conditions, the ignition delay period is short; and combustion
proceeds smoothly and rapidly, but relatively late in the cycle. By the time
combustion progresses to the main chamber, the piston is already descending.
Therefore, conditions are changing rapidly toward reduced temperatures,
thus minimizing the formation of nitrogen oxides (Ref. 6-23).
6-26
-------
GASOLINE ENGINE
DIESEL ENGINE
ROLLING
RESISTANCE
ON THE
LEVEL
1000 2000 3000 4000 5000 6000
rpm
I I I I
25
50 75
mph
100
M
140
130
120
110
100
90
80
Si 70
m 60
50
40
30
20
ROLLING
RESISTANCE
ON THE
LEVEL
1000200030004000
rpm
I I I
25
50
mph
75
Figure 6-14. Precombustion Chamber Diesel and Uncontrolled
Spark Ignition Engine Performance Maps
(Ref. 6-8)
6-27
-------
The combustion in open-chamber engines proceeds in a
different manner. For optimum smoke-free performance, combustion com-
mences prior to reaching TDC. Since the ignition delay period in these
engines is relatively long, an appreciable amount of vaporized fuel is present
in the chamber at the time of ignition. As a result, combustion proceeds
very rapidly, leading to high local flame temperatures and NO formation
X*
rates. Retarding the injection timing tends to reduce the NO emissions
X.
somewhat at the expense of a lower smoke and limited engine-power level
(Ref. 6-23).
Another important advantage of the divided chamber dies el
engine is its relatively low peak combustion pressure. Because of these
reduced pressure levels, divided-chamber engines have some weight ad-
vantage over open-chamber designs.
6.3.3 Exhaust Emissions
Although diesel engines as a group are low polluters, they
have a poor public image because of the odor and visible smoke emitted by
some diesels under certain operating conditions. In addition to odor and
smoke, diesel exhaust emissions include hydrocarbons (HC), carbon monox-
ide (CO), oxides of nitrogen (NO ), oxides of sulfur (SO ), and certain oxy-
x x
genates including aldehydes and acrolein.
The HC emissions are primarily the result of incomplete com-
bustion of fuel droplets in the cool regions adjacent to the cylinder walls and
in zones of poor mixing and, to a lesser extent, due to combustion of lubri-
cating oil. In general, the HC species found in diesel exhaust cover a wide
range of molecular weights, including species that are heavier than those
contained in the fuel. To a large extent, the level of HC emissions is related
to the design and operational characteristics of the fuel injection system and
combustion chamber. With good spray atomization and mixing, HC emissions
are generally very low.
In the main, the CO emissions from diesel engines are very low
because of the large quantity of excess air available for combustion under all
6-28
-------
operating conditions . Therefore, the level of CO emitted is related primarily
to the degree of fuel-air mixing achieved in the cylinders.
NOx represents the major pollutant specie emitted from diesel
engines. The amount of NO formed in the combustion process is kinetically
controlled and depends upon the peak combustion temperature, local oxygen
and nitrogen concentrations and the residence time of the combustion products
in the chamber. Also, fuel-bound nitrogen can have a significant effect on
the total NO emissions.
.X
Oxides of sulfur emissions are directly related to the sulfur
content in the fuel. Control of SO is achieved by limiting the amount of
X.
sulfur in the fuel.
Normally, engine blowby represents a small contribution to
the overall HC emitted from diesel engines. Control of blowby products in
naturally aspirated diesels can be achieved by positive crankcase ventilation
(Ref. 6-24).
Evaporative losses in diesels are very small because of the
low volatility of diesel fuels. Also, since the injection system is a sealed
unit, no fuel can escape through the injection nozzle following engine shutdown.
6.3.3.1 Heavy-duty Diesel Engine Emission Characteristics
Representative HC, CO, and NO specific mass emissions in
^t
terms of gram per brake horsepower-hour of heavy-duty, naturally aspirated
and turbocharged open-chamber and prechamber diesels are presented in
Table 6-2 for rated power and speed. These data represent average values
and are based on tests conducted on a total of fourteen different engines.
The HC emissions listed for turbocharged open-chamber
(direct injection) engines are higher than for naturally-aspirated engines of
this type, although one would expect the opposite. The HC emissions of
divided-chamber engines are especially low because of the high wall tempera-
ture of the prechamber and the high flow turbulence.
6-29
-------
Table 6-2. Heavy-Duty Diesel Engine Emissions
at Rated Power (Ref. 6-13)
Open Chamber, naturally aspirated
Open Chamber, turbocharged
Divided Chamber, turbocharged
*
Divided Chamber, naturally aspirated
Mass Emissions
gm/bhp-hr)
HC
1.1
1.5
0. 1
0.1
CO
7.2
4.0
1.5
3.3
NOX
7.5
10.6
5.3
4.0
Mercedes Benz 220D light-duty diesel engine (Ref. 6-25)
The variations of CO with engine type are much smaller which
is to be expected since CO is determined primarily by air-fuel ratio. Again,
the prechamber engine type is the lowest emitter.
The NO emissions of turbocharged open-chamber engines are
X.
higher than for the other types and are comparable to spark ignition engines.
This is primarily due to the high combustion temperatures achieved in turbo-
charged engines. Incorporation of an aftercooler tends to reduce the NO
emissions for turbocharged engines.
The part-load HC, CO, and NO mass emissions of diesel
jC
engines operated at rated speed are presented in Figure 6-15 in terms of
the ratio of the emissions at part load to the full-load emissions (Ref. 6-13).
As illustrated, the HC emissions increase with decreasing engine load for
all three engine types. The CO emissions of turbocharged divided-chamber
engines increase steadily when engine load is reduced, while the naturally-
aspirated and the turbocharged open-chamber diesels show an initial
6-30
-------
6 —
o
X
o
<£
3
£2
o
UJ
o
\ ^-NATURALLY ASPIRATED1
\\ ^TURBOCHARGED*
TURBOCHARGED;
a . Hydrocarbon emission, steady
state part-load , rated speed.
t OPEN
CHAMBER
* DIVIDED
CHAMBER
20 40 60 80
PERCENT OF DESIGN LOAD
4.0
b. Carbon monoxide emission, -~ z
steady state part-load, rated o^
/ iF> v~>
speed. ^?s
o
-------
reduction followed by an increase. The NO emissions of the divided-chamber
diesel increase steadily with decreasing load. In naturally-aspirated open-
chamber engines, the NO remains essentially constant at part load and
decreases for the turbocharged open-chamber configuration.
In general, slightly lower specific mass emissions are
obtained by simultaneously reducing engine speed and load.
6.3.3.2 Light Duty Diesel Vehicle Gaseous Emissions
A number of emission tests have been conducted on several
diesel engine powered vehicles by the EPA (Refs. 6-26 through 6-28), South-
west Research Institute (SWRI, Ref. 6-29), Daimler-Benz (Ref. 6-12), and
Peugeot (Ref. 6-30). To provide a direct comparison of the emissions from
equivalent diesel and spark ignition automobiles, SWRI has also tested a
Mercedes 220 gasoline engine powered car.
The emission test data presented next are based on the fol-
lowing four diesel-powered light-duty vehicles:
Mercedes Benz 220D sedan (EPA, SWRI, Daimler-Benz)
Opel Rekord 2100D sedan (EPA)
Peugeot 504 Diesel sedan (EPA, Peugeot)
Ford pickup truck with Nissan Diesel (EPA)
The Mercedes Benz 220D is a four-door sedan, powered by an
overhead cam, 134-CID four-cylinder, four-stroke precombustion chamber
diesel rated at 65 horsepower (SAE) and 4200 rpm. The vehicles tested by
the EPA and SWRI were equipped with automatic transmission and the dyna-
mometer weight was set at 3,500 pounds.
The Opel Rekord 2100D tested by the EPA was a four-door sedan,
powered by a water-cooled, four-cylinder, four-stroke diesel engine utilizing
Ricardo swirl chambers (Ref. 6-31). The engine has a displacement of 128 CID
and develops 68 horsepower (SAE) at 4200 rpm. The test vehicle had a three-
speed automatic transmission and was tested with a 3,000-pound inertia weight.
6.32
-------
The Peugeot 504 diesel tested by the EPA was a four-door sedan
with a water-cooled, four-cylinder, four-stroke, 128 CID swirl chamber
engine rated at 65 horsepower (DIN) and 4,200 rpm, and a four-speed manual
transmission. The tests were conducted with a 3, 000-pound inertia weight.
The diesel-powered, light-duty truck tested by the EPA was a
1973 model year F-250 vehicle, retrofitted with a Chrysler/Nissan four-
stroke, six-cylinder, 198-CID swirl-chamber diesel engine. The vehicle
had a four-speed manual transmission and was tested with a 4, 500-pound
inertia weight.
The available HC, CO, and NO emission data for light-duty
X.
diesel-powered vehicles tested over the Federal Driving Cycle are presented
in Table 6-3. The data represent average values from a series of runs con-
ducted on the various vehicles. For comparison, the emissions from the
gasoline engine powered Mercedes Benz 220 automobile, as tested by SWRI,
are 2.7 gm/miHC, 32.3 gm/mi CO, and 3.6 gm/mi NO .
X.
As indicated, the Mercedes Benz 22OD and the Opel Rekord
2100D meet the 1975 statutory emission standards. On the Mercedes 220D,
the average HC and CO emissions are about 50 percent below the 1975 statu-
tory standards. The average NO is approximately 50 percent below the
X.
1975 statutory standard, but four times the statutory limit for 1976. The
Daimler-Benz data are in reasonable agreement with the EPA and SWRI values.
Modifications to the fuel injection system of this vehicle resulted in a 25 per-
cent reduction of HC and CO, with no change in NO .
On the Opel, the CO and NO emissions are about 50 percent
of the 1975 statutory level. However, the average HC is very close to this
standard, with two tests indicating higher values and two indicating lower
values. The average NO is about three times the 1976 statutory limit.
The average HC emissions from the Peugeot 504 diesel, as
tested by the EPA, exceeded the 1975 standards. However, Peugeot states
that the emissions could be substantially reduced by modifying the fuel injec-
tion equipment (Ref. 6-30). The average CO emissions exceed the 1975-76
statutory standard very slightly, while the NOx emissions are about
6-33
-------
Table 6-3. Average Light-Duty Diesel Vehicle Emissions
(1975 Federal Test Procedure)
Test Vehicle
Mercedes Benz 220D
Mercedes Benz 220D
(Modified Injection)
Mercedes Benz 220D
Mercedes Benz 220D
Opel Rekord 2100D
Peugeot 504 Diesel
Peugeot 504 Diesel
Ford/Nissan Diesel
Test
Laboratory
EPA
EPA
SWRI
Daimler -
Benz
EPA
EPA
Peugeot
EPA
Test
Procedure
1975 FTP
1975 FTP
1972 FTP
1975 FTP
1975 FTP
1975 FTP
1975 FTP
1975 FTP
Emissions, gm/mi
HC
0.34
0.28
0.87
0.15
0.40
3.11
2.50
1.70
CO
1.42
1.08
1.62
2.3
1.16
3.42
2.20
3.81
NOX
1.43
1.48
1.83
1.45
1.34
1.07
1.30
1.71
Reference
6-28
6-28
6-29
6-12
6-28
6-28
6-30
6-27
30 percent below the 1975 standard. The HC and CO data provided by Peugeot
are substantially lower than the EPA data, while NO is slightly higher.
Jt.
The modified Ford pickup truck with the Nissan diesel engine
shows average NO emissions approximately 45 percent below the 1975
standard. The CO emissions are about 10 percent above the standard, while
HC is about four times higher than the 1975-76 standard.
Emission data taken by Ricardo Engineers on a number of
European diesel-powered automobiles tested in accordance with the 1975
Federal Driving Cycle are listed in Table 6-4. These data are in reasonable
6.34
-------
Table 6-4. Diesel Automobile Emissions (Ref. 6-32)
Fuel Injection System
Standard
Optimized Injection Timing
Optimized Injection Timing + EGR
Emissions,' gm/mi
HC
0. 17
0. 17
0.17
CO
1.6
1.6
2.5
NO
X
1.2
0.7
0.4
1975 Federal Test Procedure
agreement with the previously discussed data. Also presented in the table
are projected emissions to be achieved by optimized timing and EGR, at
some increase in fuel consumption. Ricardo estimates that achievement
of the 0.7 gm/mi NO level would result in a 10 to 12 percent loss in fuel
X.
economy.
When evaluating these emissions, one must consider the small
data sample available for each vehicle type. Therefore, an assessment of
car-to-car variability and the effects of mileage accumulation on emissions
is not possible. Also, it should be noted that the engine power of some of
the vehicles was insufficient to negotiate the first and second acceleration
modes of the Federal Driving Cycle.
6.3.3.3 Smoke and Particulate Emissions
The smoke emitted by dies el-powered vehicles can be grouped
into three categories -- white, blue, and black. White smoke or cold smoke,
which usually appears in the exhaust during an engine cold start, consists of
partially oxidized fuel droplets combined with other compounds, such as
aldehydes. It is formed in the wall region of the cylinders where the
temperatures are not high enough to ignite the fuel. Blue smoke is emitted
6-35
-------
by some diesels as a result of the combustion of excessive amounts of
lubricating oil. Black smoke, or hot smoke, consists of agglomerated car-
bon particles and partly burned hydrocarbon species formed as a result of
incomplete combustion in fuel-rich regions of the combustion chamber
(Ref. 6-2).
Smoke data from four naturally aspirated, four-stroke diesel
engines are depicted in Figure 6-16 . The sharp increase of the smoke level
of engines A and B at high loads may be caused by poor injector performance,
resulting in inadequate fuel atomization and fuel impingement on the cylinder
walls. In general, turbocharged engines exhibit lower smoke levels at full
load because of the higher excess air available (Ref. 6-25). Springer and
Ashby conducted smoke tests on a diesel-powered Mercedes Benz 220D auto-
mobile in accordance with the Federal Smoke Test Procedure for heavy-duty
vehicles (Ref. 6-33). The results from these tests are presented in Table 6-5,
showing smoke opacity during low gear acceleration ("a"-factor), lug-down
(" b"-factor), and at peak smoke conditions ("c"-factor). As indicated, the
engine meets the 1974 federal heavy-duty smoke standards.
Particulate emissions from the Mercedes 220D engine
determined by the Dow Chemical Procedure are listed in Table 6-6. The
average particulate emission of 0.73 g/mi is about 2 to 3 times higher than
for a gasoline engine using leaded fuel and about 10 to 20 times higher than
for nonleaded gasoline. The material collected was pitch black and very
fine (Ref. 6-25). Tests conducted by Daimler-Benz indicate average particu-
late emissions of 0.45 g/mi (Ref. 6-12).
Smoke and particulate emissions from diesel engines can be
minimized by several means, including proper injector design and maintenance,
engine derating, smoke suppressant additives in the fuel, and possibly filters
and particulate traps (Ref. 6-34). The additives can be divided into two
groups. One group is represented by the detergent types which maintain a
clean fuel injection system. The other group consists of metal-containing
materials which reduce the ignition temperature of the carbon and promote
its oxidation.
6-36
-------
30 r-
25 50 75
RATED POWER, percent
Figure 6-16.
Exhaust Smoke vs Output Power for Four Open-Chamber,
Naturally Aspirated, Heavy-Duty Diesel Engines
(Ref. 6-25)
6-37
-------
Table 6-5. Smoke Emissions from Mercedes Benz 220D
Automobile, Simulated Federal Smoke Tests
(Ref. 6-33)
Average from 4 runs
Federal Heavy Duty Limit, 1970
Federal Heavy Duty Limit, 1974
Percent Opacity
"a" factor
10.8
20
15
"b" factor
9.4
40
20
"c" factor
15.1
-
50
Table 6-6. Particulate Emissions from Mercedes Benz
220D Automobile, Dow Chemical Procedure,
gm/mile (Ref. 6-33)
Sample Flow,
cfm
4
1
1
1
Filter Type
Fiberglass
Fiberglass
Fiberglass
Anderson + Millipore
Average
Mass, gpm
0.66
0.80
0.75
0.73
0.73
6-38
-------
Some work with fuel additives containing barium has been
conducted by several investigators (Refs. 6-35 and 6-36). In all cases, black
diesel smoke was substantially reduced with relatively small amounts of
additives. However, there may be health effects resulting from such addi-
tives, and they are currently not recommended for use.
6.3.3.4 Emission Control Techniques
A number of emission control techniques and devices have
been considered by many investigators for potential use in diesel engines.
These include injection timing retardation, valve overlap, EGR, water
injection, catalysts, thermal reactors, and turbocharging. These approaches
are briefly described in the following paragraphs.
Injection timing retardation in general has a beneficial effect
on NO emissions at the expense of higher HC and CO and reduced fuel
X.
economy. However, since HC and CO emissions are inherently low in diesels,
a tradeoff between these species might be feasible (Refs. 6-23, 6-37, and
6-38).
Modification of the valve overlap can reduce the NO emissions
^ x
by reducing the degree of engine scavenging. However, the benefits derived
from this approach are apparently rather limited (Refs. 6-12 and 6-24).
Exhaust gas recirculation (EGR) appears to be the most promis-
ing technique for diesel NO control. As in spark ignition engines, EGR
.JC
acts as a charge diluent and reduces the peak temperature in the combustion
chamber. Normally, this results in some increase in fuel consumption and
HC, CO, and smoke emissions (Refs. 6-2, 6-23, and 6-37). Conversely,
test data from a Mercedes Benz 220D vehicle presented in Figure 6-17 indi-
cate no change in HC, CO, and smoke for EGR rates up to about 25 percent
(Ref. 6-12). Utilization of 25 percent EGR has resulted in a 25 to 50 percent
reduction in NO , depending upon the operating conditions of the engine. Simi-
X.
lar results have been reported by Marshall and Fleming (Ref. 6-37). According
to Daimler-Benz, there is apparently no loss in part load fuel economy with
EGR (Ref. 6-12). However, deposit buildup and corrosion in the intake ports
6-39
-------
2400 rpm
26.5 psi BMEP
SMOKE SZ-BOSCH SCALE
2400 rpm
53 psi BMEP
SMOKE. SZ-BOSCH SCALE
J
1
CO
0.2
Vol-%
0
160
ppm-C,
80
40
0
NO*
200
ppm
100
50
0
«
_.^
^
'
'
/
S
f
—
•- =i
^^
^s
X
KNOCKING-^
Jl
10 20 30 %
EGR
j
1
CO
Op
Vol-%
0
HC
120
ppm-C,
40
0
NO.
PPm
300
200
100
0
/
/
J
\
^
_KN
OCKIN
Y£
~ f&
10 20 % 3(
EGR
Figure 6-17. Effect of Exhaust Gas Recirculation on the Emissions
from a Light-Duty Diesel Engine (Ref. 6-12)
have been experienced by Daimler-Benz which might create a valve durability
problem. Caterpillar has encountered similar problems (Ref. 6-24).
As with EGR, water injection into the engine intake is an ef-
fective means to reduce NO emissions from diesel engines (Refs. 6-2, 6-29,
J\.
and 6-36). This is illustrated in Figure 6-18 showing NO emissions, from
Ji.
a turbocharged heavy-duty diesel engine operated at 2,200 rpm, as a function
of water-injection rate. Since the inducted water vaporizes and displaces air,
the air flow rate into the engine decreases, and this may result in an increase
in smoke. Potential problem areas related to water injection include corro-
sion of the induction system and the intake valves.
The HC and CO emissions from dies els can be further reduced
by means of catalytic converters installed in the engine exhaust system
(Ref. 6-2). However, in view of the already low HC and CO emissions of
diesel engines, this approach is probably not very cost-effective. Control of
6-40
-------
1400r-
ENGINE SPEED:
2200 rpm
60% RATED
TORQUE
100% RATED TORQUE
1.0 1.5 2.0 2.5
Ib OF WATER/lb OF FUEL
3.0
Figure 6-18 .
Effect of Water Induction Rate on the NO Emissions
of a Turbocharged Diesel Engine (Ref. 6-24)
6-41
-------
NO bv means of reducing catalysts is not possible for diesels because of
x 7
the excess oxygen present in the exhaust.
Thermal reactors, although quite effective in reducing HC and
CO from spark ignition engines, are not considered feasible for diesels because
of the relatively low exhaust temperature and HC and CO concentrations.
Turbocharging represents an effective technique to reduce
HC and CO emissions, and incorporation of an aftercooler tends to reduce
the NO emissions as well (Ref. 6-2 and Ref. 6-23).
x
6.3.4 Fuel Economy
Because of their superior fuel consumption characteristics,
combined with high reliability and durability, dies el engines have long been
favored for use in many truck and industrial applications. This is illustrated
in Figure 6-19 , showing the steady-state fuel economy of 1972 model year
diesel and gasoline engine powered Mercedes 220D and 220 vehicles as a
function of level-road speed (Ref. 6-39). As shown, the fuel economy of the
diesel vehicle at 30 mph is about 50 mpg, compared with about 25 mpg for
the gasoline vehicle adjusted to meet the 1972 Federal emission standards.
With increasing road speed, the advantage of the diesel with respect to fuel
economy diminishes. It is evident that the fuel savings of a diesel, compared
with a gasoline engine, are strongly affected by the vehicle driving cycle.
The fuel savings are negligible at freeway speeds, about 30 percent over
country roads, and about 50 percent in taxi operation (Ref. 6-32). At idle,
the fuel savings are of the order of 70 percent.
Fuel economy data for the Mercedes Benz 220D, Opel Rekord
2100D, Peugeot 504 diesel, and Ford/Nissan diesel vehicles, as tested by
the EPA in accordance with the 1975 and 1972 Federal Test Procedures, are
presented in Table 6-7 (Refs . 6-26 through 6-28). The figures were com-
puted by the EPA from the measured HC, CO, and CO exhaust concentra-
tions using the carbon balance method. Also listed in the table are the fuel
economy values for the 1973 model year gasoline engine powered Mercedes
Benz 220, the Peugeot 504, and the average 3, 000-pound, 3, 500-pound and
4500-pound 1973 automobiles as determined by the EPA from certification
6-42
-------
o
I
60
50
o> 40
Q.
30
O
o
LU
D 20
10
MODELrMB 220 (134.0 cu in.)
, MODEL YEAR
MODELrMB 220D (134.0 cu in.) DIESEL
MODELrMB 220 (134.0 cu in.) GASOLINE, EUROPE
10
20
30 40 50 60
VEHICLE ROAD SPEED, mph
70
80
90
Figure 6-19. Fuel Economy versus Vehicle Speed for Diesel and Gasoline Automobiles
(Ref. 6-39)
-------
Table 6-7. Light-Duty Diesel and Gasoline Vehicle Fuel Economy, miles/gallon
(Refs. 6-26, 6-27, 6-28)
1975 FTP
Test No.
Mercedes
220D
Std. Inj.
(3,500)*
1 23.7
2
3
4
5
Average
22.6
23.8
24.0
24. 0
23.6
Mercedes
220D
Mod. Inj .
(3,500);:
23.8
24.4
25. 3
24.9
24.5
24.6
Opel
2100D
(3,000)*
23.9
24.3
23. 2
23.7
-
23.8
Peugeot
504
Diesel
( )**
25.4
25.2
24. 1
25.2
26.2
25.2
Ford/Nissan
Pick- Up
(4,500)*
22. 2
21.3
20.7
-
-
21.4
Mercedes
220D
Std. Inj.
(3,500)*
23.3
22.3
33.3
23.9
23.5
23. 3
1972 FTP
Mercedes
220D
Mod. Inj.
(3,500)*
22.8
23.8
24. 0
24.0
23.6
23.6
Opel
2100D
(3,000)*
23.3
23.5
22.7
23. 1
-
23.2
Peugeot
504
Diesel
( )**
24.3
24.3
23.3
24.2
25. 1
24.2
Comparative Gasoline Powered Cars
1973 Mercedes Benz 220, Gasoline Car 13.2
1973 Peugeot 504, Gasoline Car 17.0
1973 Average 3000-lb Gasoline Car 15.6
1973 Average 3500- Ib Gasoline Car 13.9
1973 Average 4500- Ib Gasoline Car 10. 1
Test inertia weight, pounds
Inertia weight not available
-------
data (Ref. 6-40). As indicated, the fuel economy of the diesel-powered
vehicles is substantially higher than that of equivalent gasoline automobiles.
For instance, the Mercedes Benz 220D diesel with standard fuel injection has
a fuel economy of 23.6 mpg, compared with 13.2 mpg for the gasoline engine
powered Mercedes 220 with automatic transmission and 13.9 mpg for the aver-
age 3,500-pound 1973 model year certification vehicle. These differences cor-
respond to fuel savings of about 41 percent and 44 percent, respectively. The
fuel economy values for the Opel 2100D and Peugeot 504 diesel automobiles are
23.8 mpg and 25.2 mpg, respectively, compared with 15.6 mpg for the average
3000-pound 1973 model year car. The corresponding fuel savings are about
42 percent and 38 percent, respectively. For the Ford/Nissan diesel powered
pickup truck, the measured average fuel economy is 21.4 mpg, which is more
than twice that of the average 4,500-pound 1973 model year automobile.
However, when comparing the fuel economy of current diesel
and gasoline powered automobiles, an adjustment must be made to account
for the higher power and torque output of the gasoline engine. This is illus-
trated in Figure 6-20, showing the fuel economy of 3,000-pound and 3,500-pound
1973 model year certification vehicles as a function of rated engine power.
Since the fuel economy increases with decreasing power, the fuel savings
realized with the diesel over gasoline engines of comparable performance
would be somewhat lower than noted above.
6.3.5 Noise
In general, diesel engines are inherently noisier than gasoline
engines, in particular the open-chamber designs most frequently used in truck
and industrial applications. The noise is especially strong at idle and during
a cold start of the engine. Although the noise levels of light-duty automotive
diesel vehicles are generally considered acceptable in the normal operating
regime of the engine, the noise from these engines is noticeably higher than
that of modern V-8 gasoline engines (Refs . 6-12, 6-41, and 6-42).
The principal contributors to diesel noise are the combustion
pressure transients, piston slap, and timing gear impacts. Other sources
include valve gear and fuel injection system impacts (Ref. 6-42).
6-45
-------
u
LJ
20
18
E 16
I
UJ
12
10
8
MEAN TREND LINE
3000-Ib CARS
O
O 3000-Ib INERTIA WEIGHT
D 3500-lb INERTIA WEIGHT
AVERAGE FUEL ECONOMY OF ALL 3000-Ib CAR
AVERAGE FUEL ECONOMY
OF ALL 3500-lb CARS
D
D
D
D
100
150 200
ENGINE BRAKE HORSE POWER, bhp
250
Figure 6 -20. Fuel Economy versus Engine Rated Power,
1973 Model Year Certification Vehicles
-------
Exterior and interior noise levels reported by Daimler-Benz
for its Mercedes 220D diesel and Mercedes 220 gasoline vehicles are shown
in Table 6-8 and Figure 6-21. With the exception of a few operating points,
the diesel noise is of the order of 2 to 3 dB(A) higher than the noise of the
gasoline engine, as shown.
Interior noise data from Peugeot 504 diesel and gasoline engine
powered automobiles are listed in Table 6-9 for a number of steady-state
operating conditions of the vehicles. Again, the diesel engine tends to be
slightly noisier than the gasoline engine (Ref. 6-30).
Several approaches are currently being considered by industry
to reduce the noise of diesel engines. These include reduction of component
clearances and structural changes on the engine and engine compartment to
modify their vibration characteristics (Refs. 6-12 and 6-42). Intake air heat-
ing might prove effective to reduce the noise at idle.
6.3.6
Odor
Diesel odor has long been recognized as a very undesirable
exhaust emission product. Although a considerable amount of research has
Table 6-8. Exterior Vehicle Noise, Mercedes Benz 220D Diesel
and 220 Gasoline Automobiles (Ref. 6-12)
Engine Start
Idle
Normal Vehicle Start
Fast Vehicle Start
30 mph (third gear)
44 mph (fourth gear)
62 mph (fourth gear)
Exterior Noise, dB(A)
Mercedes 220D
65
56
72
78
75
83
79
Mercedes 220
60
51
69
80
70
81
75
6-47
-------
CD
TJ
80
70
60
80
70
60
50
ACCELERATION
DECELERATION
220D
220
1000
2000 3000 4000
ENGINE SPEED, rpm
5000
6000
Figure 6-21. Interior Vehicle Noise, Mercedes Benz 220D Diesel
and 220 Gasoline Automobiles (Ref. 6-12)
Table 6-9. Interior Noise, Peugeot 504 Diesel and Gasoline
Automobiles (Ref. 6-30)
Speed
| Idle
; 31 mph
50 mph
62 mph
74 mph
87 mph
Maximum Speed
Peugeot 504 v
Gasoline Sedan
dB(A)
50
65
70
72
76
79
82
Peugeot 504 *
Diesel Sedan
dB(A)
54
67
70
72
78
-
-
Noise measurements taken in the interior of the vehicle
at constant speed in fourth gear in dB(A)
6-48
-------
been under way for some time, progress toward determination of the cause
of the odor has been very slow because of the complexity of the heterogeneous
combustion occurring in diesels, combined with the lack of accurate instru-
mentation (Refs. 6-43 and 6-44).
The diesel odor is known to be related to the combustion
process, chamber shape, and (to a lesser degree) fuel type and fuel
composition.
Diesel exhaust contains small quantities of a large number of
unburned and partially burned hydrocarbon compounds. Some of these com-
pounds have strong odors, even in very low concentrations, and mixtures
of these materials are believed to be responsible for the formation of the
typical diesel odor (Ref. 6-2). Two major types of odor have been identified,
one being an "oily-kerosene" type apparently caused by aromatic hydrocarbons
and the other being of the " smokey-burnt" type, and probably caused by par-
tially oxidized compounds. Barnes (Ref. 6-45) has concluded that diesel odor
is formed by partial oxidation reactions in the fuel-lean regions of the engine.
In the absence of reliable instrumentation, odor has been
evaluated in the past by specially selected human panels trained to recognize
both odor quality and intensity. This panel classifies the odor by comparing
it with 28 different odor qualities and intensities supplied from the bottles of
the Quality/Intensity Evaluation kit. The overall diesel composite rating "D"
ranges from D-l to D-12, with D-12 representing the strongest odor. Burnt-
smokey quality "B, " aromatic quality "A, " oily quality "O, " and pungent
quality "P" have a range between 1 and 4. More recently, a diesel odor
analysis instrument has been developed by A. D. Little, Incorporated, which
is designed to permit direct measurement of diesel odor levels (Ref. 6-46).
Evaluation of the instrument is in progress.
The effect of engine operating condition on odor intensity can
vary markedly for different engines. For naturally aspirated engines run at
rated speed, the minimum odor intensity seems to occur in the mid-power
range of the engine; in turbocharged engines, odor intensity remains essen-
tially constant over the power range. This trend is believed to be the result
6-49
-------
of a reduction in the ignition delay time occurring with rising power output
due to an increase in the intake and compression temperatures. Since odor-
ants are thought to be formed by partial oxidation of the fuel prior to ignition,
the constant odor intensity pattern would follow from this effect (Ref. 6-35).
Odor data reported by Springer for a Mercedes Benz 220D
diesel vehicle are listed in Table 6-10 (Ref. 6-29). The data show no clear
trend of odor intensity with engine speed and load. The highest odor intensity
was recorded during engine deceleration. In general, the odor levels of the
Mercedes engine are quite comparable to the levels observed on heavy-
duty diesel engines, although the character of the odor appears somewhat
different. Other exhaust emission species, including HC, CO, NO , aide-
.X
hydes, formaldehyde, and acrolein, measured along with the odor, are listed
in Table 6-10. As indicated in the table, odor intensity in the diesel exhaust
correlates reasonably well with HC and acrolein emission concentrations. It
should be noted, however, that the mass emissions of the aldehydes, formal-
dehyde, and acrolein were much lower for the Mercedes Benz 220D diesel
than for the 220 gasoline engine. These results are difficult to explain and
require further study.
Several odor-abatement approaches have been considered by
industry, including injector modifications, catalytic mufflers, and fuel addi-
tives. Utilization of a needle-valve injector with a very small dribble volume
in place of a check-valve-type injector has resulted in a substantial reduction
of the odor intensity from two General Motors two-cycle, GV-7 1 diesels. Also,
catalytic mufflers have shown some benefit, but in early tests their effective-
ness deteriorated rapidly with mileage accumulation. However, in view of
the catalyst improvements achieved in the past few years, further work in the
area of catalytic odor abatement might be desirable. With respect to fuel
additives, some reduction in odor intensity has been demonstrated by South-
west Research Institute (Ref. 6-47). However, the benefits achieved to date
are small and more research would be required to permit a meaningful fea-
sibility assessment of this technique.
6-50
-------
Table 6-10. Average Odor and Gaseous Emissions from Mercedes Benz
220D Automobile (Ref. 6-29)
Engine Speed
RPM
2500
2500
2500
4200
4200
4200
Idle
Acceleration
Deceleration
Load
Percent
0
50
100
0
50
100
0
-
-
Odor
Composite
"D"
6.0
3.8
4.0
4.4
6.1
5. 5
3.9
5.2
6.8
Burnt
"B"
2.0
1. 1
1.2
1.3
2. 1
1.9
1. 1
1.8
2.1
Oily
"O"
1.4
1. 0
1.0
1.1
1.3
1.3
1.0
1.2
1.7
Atomatic
"A"
1.1
0.8
0.9
1.0
1. 1
1.1
1.0
1.0
1.2
Pungent
iipn
1.5
0.7
0.8
1.0
1.6
1.3
0.8
1.2
1.8
Gaseous Emissions, oom
HC
980
71
64
89
739
60
105
-
-
CO
7.3
231
209
348
1155
968
168
-
-
NOX
78
339
505
224
528
598
124
-
-
Acrolein
7.7
1.4
1.0
1.2
6.3
1.5
1.6
-
-
Formal-
dehyde
15.4
15.2
13.7
7.9
20.0
21.7
12.0
-
-
Aliphati
Aldehyde
28.2
2L1
14. 1
12.6
35. 1
25. 1
19.0
-
-
-------
6.3.7 Reliability and Durability
Because of a lack of applicable test data, a quantitative
evaluation of the reliability and durability characteristics of light-duty auto-
motive diesel engines is not possible. However, in general terms, the
reliability of diesels is quite high and probably superior to equivalent gaso-
line engines (Ref. 6-12). The excellent reliability of heavy-duty automotive
and industrial diesels is well known.
Although engine durability depends on many parameters (such
as individual driving patterns; maintenance programs; and,to a lesser extent,
fuel quality) there are indications that light-duty automotive diesels are more
durable than equivalent gasoline engines. In truck and industrial applications,
the ruggedness and superior durability characteristics of the diesel engine
have been adequately demonstrated. Many truck diesels have been operated
over 300, 000 miles and more with minimum maintenance.
6.3.8 Maintenance
With respect to engine maintenance, the diesel engine has
definite advantages over the gasoline engine simply because it has fewer parts
that require periodic adjustment. For instance, the diesel has no carburetor
and ignition system; hence tuneups and replacement of spark plugs, points,
and condenser are not needed. This results in substantial cost savings.
The scheduled maintenance on diesels includes periodic valve
adjustments, oil and oil filter changes, and inspections and adjustments of
the fuel injection system. According to Daimler-Benz, the injection system
in its 220D automobile is practically maintenance-free (Ref. 6-12). Gener-
ally, maladjustment of the fuel injection system or carbon buildup on the
nozzles is accompanied by the appearance of visible smoke in the exhaust
(along with an increase in HC and CO emissions) and, in severe cases, by a
noticeable loss in power. As a result, it is more likely that the owner of a
diesel vehicle would take the necessary steps to eliminate the problem than
the owner of a gasoline car where quite frequently, fouling of certain
6.52
-------
components might greatly increase the emissions without affecting
engine performance.
6.3.9 Safety
Since dies el fuel is much less volatile than gasoline, formation
of explosive fuel air mixtures in enclosed areas due to fuel evaporation is
essentially impossible. Also, in the event of an accident, the chance of an
engine or grass fire resulting from spilled fuel is extremely remote. There-
fore, diesel is considered to be a safer fuel than gasoline.
6.3.10 Driveability
A number of performance parameters have been identified for
gasoline engine powered automobiles having an adverse effect on vehicle
driveability. These include (1) engine stalling during start, (2) engine
stumble, hesitation, surging, knocking, and backfire during acceleration, and
(3) backfire and stalling during vehicle deceleration. In diesel engines, defi-
ciencies of this type are not common; hence, driveability of diesel-powered
automobiles should be excellent.
With respect to acceleration, available test data indicate that
current diesel-powered automobiles accelerate much more slowly than gaso-
line vehicles of similar weight and engine size. For example, the Mercedes
Benz type 220D automobile requires 41.5 seconds to accelerate from 30 mph
to 70 mph, compared with 18.3 seconds for the Mercedes Benz 230 gasoline
car adjusted to meet the 1975 Federal emission standards, and 12 seconds
specified for the U.S. Department of Transportation's Experimental Safety
Vehicle (Ref. 6-12). However, the acceleration capability of diesel vehicles
could be substantially improved by supercharging.
6.3.11 Cost
The diesel-powered automobiles currently marketed in the
United States by Daimler-Benz and in Europe by Daimler-Benz, Peugeot, and
Opel are more expensive than their spark ignition counterparts. According
6-53
-------
to Daimler-Benz (Ref. 6-12), the cost differential is due to a number of
unique additional components required in diesel engines, including the fuel
injection system,pump drive, cylinder liners, special pistons, and glow
plugs. In addition, the sturdier construction, the complex prechamber, and
the larger starter motor and battery required in the diesel further contribute
to the higher cost of this engine.
To illustrate this point, the purchase price of the diesel-
powered Mercedes Benz 220D is currently $211 higher than the otherwise
identical Mercedes 220 gasoline car. However, the higher initial cost of the
diesel is counteracted by the lower fuel consumption and maintenance cost
of this engine. Furthermore, the cost differential diminishes as more emis-
sion control equipment is added to the gasoline vehicle to meet future emis-
sion control standards.
According to Peugeot and Opel, the cost differential in Europe
between diesel and gasoline engine powered automobiles is approximately
$600 to $800. In California, the base price of the Peugeot 504 diesel sedan
is $5800 as compared to $4830 for the standard 504 vehicle with a spark-
ignition engine.
In the case of Daimler-Benz, the relatively small cost differ-
ential between the 220 series diesel and gasoline vehicles is due to the fact
that the two engines were designed with commonality of parts as a prime con-
sideration. Conversely, diesels manufactured by Peugeot and Opel are
special designs, and their higher cost reflects the low-production volume of
these engines.
Currently, no reliable data are available regarding the effects
of turbocharging on the cost of light-duty diesel engines. In truck engines,
the cost of the turbocharger is of the order of 15 percent of the cost of the
engine. Based on a preliminary design study, Ricardo feels that a new light-
duty diesel engine could be designed which would be acceptable to the U.S.
consumer when installed in a medium-size automobile. The cost of this
100 to 130 hp engine on a mass production basis would be about 40 percent
6-54
-------
higher than that of an equivalent gasoline engine, resulting in a vehicle cost
increase of about 5 percent (Ref. 6-32).
6.4 CURRENT STATUS OF TECHNOLOGY
Diesel engines are widely used in Europe and Japan for taxi-
cabs and other light-duty vehicles, primarily because of their lower fuel
consumption. However, the power output of these engines is considerably
lower than that required for the heavier American vehicles, and the operating
conditions are generally different.
The Mercedes Benz 220D diesel engine powered automobile
has been marketed in the United States since 1956. Early in 1974, Peugeot
of France offered its model 504 diesel passenger car for sale in this
country. Conversely, General Motors (Opel of Germany) has no plans
to enter the United States market with its diesel-powered automobile
model.
Daimler-Benz of Germany has been producing diesel automobiles
since 1936. To date, over one million of these vehicles have been sold to cus-
tomers in over 100 countries. According to Daimler-Benz (Ref. 6-12), the
diesel vehicles are marketed primarily in those areas of the world where fuel
cost is high and where engine power is not a prime consideration. Between
1956 and 1972, a total of 53,331 diesel automobiles have been sold by Daimler-
Benz in the United States, with more than 6,000 sold in 1972 alone. In 1971,
Daimler-Benz manufactured over 99.000 Mercedes Benz 220Ds out of a total
car production of about 284, 000.
Peugeot introduced its diesel-powered automobile in 1959 and
has since sold a total of about 460, 000 vehicles. In 1972, diesel sales
amounted to nearly 80, 000 out of a total production of 671, 000 vehicles.
The available test data indicate that the emissions of diesel-
powered light-duty vehicles manufactured by Daimler-Benz and Opel are be-
low the statutory 1975 Federal emission standards. NOx levels of the order
of 0.8 g/mi or even lower appear achievable by means of EGR or water injec-
tion, combined with injection system modifications.
6-55
-------
The fuel economy (mpg) of the current light-duty diesel vehicles
over the Federal Driving Cycle is between 50 percent and 70 percent better than
that of the average 1973 certification vehicles tested at the same inertia weight.
6. 5 PROJECTED STATUS
In the absence of adequate technical information on advanced
light-duty, lightweight diesel engines, it is extremely difficult to make mean-
ingful projections regarding the future use of advanced diesels in passenger
cars.
According to Daimler-Benz (Ref. 6-12), the diesel engine
should not be considered as a replacement of the gasoline engine, but rather
a supplementary concept for specific applications where fuel economy and
engine durability are of prime concern. However, with the advent of more
stringent emission control requirements, and with the energy shortage in
mind, interest in the diesel as an automobile power plant is expected to
increase in the future. In spite of these developments, it will be very diffi-
cult to substantially increase the production volume of diesels within a
reasonable time span. In this regard, Daimler-Benz feels that a minimum
of 24 months would be required to increase its current production volume on
the 220D vehicle. Conversely, a lead time of 5 to 6 years is projected by
Daimler-Benz for the design and development of a new, more powerful diesel
engine and for the construction of a production facility (Ref. 6-12).
It is difficult to ascertain whether the general public in the
United States would ever be willing to accept the low power capability of the
current diesel-powered automobile. Therefore, it appears appropriate to
briefly dwell upon a number of approaches which could be pursued to increase
the power output of diesel engines and to reduce their weight and bulk.
Turbocharging or supercharging is considered to be a feasible tech-
nique to increase the specific power output of diesels with a concomitant improve-
ment in fuel economy. However, the current light-duty diesel engine designs may
not be rigid enough to withstand the higher combustion pressures and temperatures
encountered by adding a turbocharger. The stress problems in the cylinders,
6-56
-------
pistons, and crankshaft could be alleviated by reducing the compression ratio
of the engine. Since the compression temperature decreases with decreasing
compression ratio, engine starting becomes more difficult, especially at low
ambient temperatures, and starting aids would have to be utilized such as
glow plugs, intake air heaters, ether injection, preinjection of fuel, or even
variable compression ratio pistons. This is illustrated in Figure 6-22, show-
ing the compression temperature and the ignition temperature of diesel fuel
and gasoline as a function of the compression ratio. Using diesel fuel and
operating at normal ambient conditions, the diesel can maintain idle speed
for compression ratios above about 12:1 (Ref. 6-15).
In general, turbochargers are not very effective at low engine
speeds, and they have inadequate response to the load variations encountered
in light-duty vehicle operation. These deficiencies are overcome by the
Comprex supercharger, which has a much faster response and higher BMEP
capability in the low speed regime of the engine. As a result, the low speed
torque of the engine increases substantially, and this permits a reduction of
the design peak power at a concomitant savings in engine weight and bulk.
Additional weight savings might be possible by reducing the design life of the
diesel engine, which is several hundred thousand miles in current designs.
Ricardo and Company Engineers of England, under contract to
the EPA, is currently involved in a survey of the problem areas and the impact
of diesels as power plants for passenger car applications. The parameters
being considered by Ricardo include two stroke vs four stroke, naturally aspi-
rated and turbocharged diesels, rotary diesels, EGR, startability. emissions,
odor, and noise.
With regard to larger cylinder and engine displacement, it
appears that this approach alone is not sufficient to achieve an adequate
increase in engine power. Apparently, the requirement for good combustion
and smoothness of operation limits the maximum cylinder volume of high-
speed prechamber diesel engines to about 36 CID. Hence, four-cylinder
6-57
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UJ
UJ Q.
I- 2
z y
O h
o Z
ft O
jl
1§
O ~~
u
1600
1400
1200
1000
800
400
200
COMPRESSION TEMPERATURES
--O- IGNITION TEMPERATURES
FULL LOAD
—o-
COLD START
-25 F
DIESEL
AMBIENT FUEL
10
12
14 16 18 20
COMPRESSION RATIO
22
24
Figure 6-22. Compression and Ignition Temperatures
versus Compression Ratio (Ref. 6-15)
6-58
-------
engines are limited to a displacement of about 150 CID, and additional
cylinders would be required to increase the engine displacement beyond that
point. As a result, the weight and volume of the engine would increase
accordingly, unless any of the above modifications were to be incorporated.
6-59
-------
SECTION 7
-------
7. WANKEL AND OTHER ROTARY PISTON ENGINES
7.1 INTRODUCTION
7.1.1 General Description
In its simplest form, the Wankel engine consists of a single
triangular-shaped rotor which revolves eccentrically in a chamber with a
double-lobed cross section (see Figure 7-1). The action of the rotor is
such that the apexes of the rotor continuously brush the walls of the
chamber, thereby forming three compartments which rotate around the
chamber, changing size and shape as they move by virtue of the eccentricity
of the rotation. Thermodynamically, the engine is based on the Otto cycle
in its "four-stroke" form. On one side of the chamber casing is a spark
plug. Each compartment passes successively through all phases of the
Otto cycle once in each rotor revolution. Since there are three compart-
ments, three power strokes are produced with each rotor revolution.
7.1.2 Historical Development
Many efforts have been expended in the past century to
develop internal combustion engines that will produce torque at a reasonably
low speed without a crankshaft and its attendant complication. Within the
last decade or so, these efforts have produced a wide variety of prototype
rotary piston engines largely experimental in nature. Among these rotary
engine designs the one presently receiving the greatest attention for auto-
motive application is the Wankel Rotary Piston Engine invented by
Felix Wankel in 1953. The first prototype of this engine was run success-
fully at the NSU Works in Germany on February 1, 1957- Later that year,
the design was modified to its present basic form, the KKM-125 engine' ,
which is the forerunner of all current Wankel Rotary Piston Engines.
*"KKM" stands for Kreiskolbenmotor, orbiting piston engine. The "125"
designates the engine chamber volume in cubic centimeters.
7-1
-------
AIR INTAKE
EXHAUST
SPARK
PLUG
COMPRESSION
b.
c.
POWER
EXHAUST
Figure 7-1. Wankel Rotary Engine
7-2
-------
Since this early introduction, experimental engines of various
sizes have been built and tested for a wide variety of applications. The
patent is owned by NSU Motor Werke of Germany, which has begun to manu-
facture the engine in small numbers. NSU also licenses the manufacture
of the Wankel engine by a number of other major companies, including
Curtis s-Wright--and, more recently, General Motors--in the United States.
These engines range in output power from about 0. 6 hp for model airplanes
to 780 hp for a Curtis s-Wright experimental engine. Typical applications
cover automobiles, aircraft, boats, snowmobiles, lawn mowers, and
auxiliary power units.
Current automotive rotary-piston engines are water cooled
and are designed with one or two rotors for production cars and up to four
rotors for experimental models. A list of vehicles equipped with these
engines together with pertinent data is given in Table 7-1 (Ref. 7-1).
Table 7-2 summarizes manufacturers known to be currently active in rotary
engine development, together with the areas of application for the develop-
mental engines.
7.2 POWER PLANT DESCRIPTION
7.2.1 Power Plant Configuration
Figure 7-2 shows a cutaway drawing of a typical two-rotor
automotive production engine. The basic engine consists of a triangular
rotor that orbits within an outer housing. The volume enclosed between the
housing and each rotor flank varies as the rotor apex orbits, tracing an
epitrochoid surface (the housing surface). This motion produces the four
strokes of the conventional four-stroke Otto-cycle, (i.e.; intake, com-
pression, expansion, exhaust) for each revolution of the rotor (Figure 7-3).
Intake and exhaust processes occur through ports instead of valves, thus
eliminating the entire valve train of the reciprocating engine which simplifies
the basic engine design. The intake and exhaust flows are controlled by the
7-3
-------
Table 7-1. Vehicles Equipped with Rotary
Piston Engines (Ref. 7-1)
Vehicle
NSU Wankel
Spider
NSU Ro-80
Mazda Cosmo
1 10-S
Mazda R-100
Mazda RX-2
Citroen M-3 5
Mercedes Benz
C-1U MKI
Mercedes Benz
c-m MK ii
Mustang RC 2-60
Introduction
1964
1967
1967
1969
1970
1970
1969
1970
1965
Discontinued
1967
—
1969
—
—
Limit 500
cars
Experimental
Experimental
Experimental
Engine
NSU
NSU
Toyo Kogyo
Toyo Kogyo
Toyo Kogyo
NSU
Mercedes Benz
Mercedes Benz
Curtiss -Wright
Horse-
power
64
136
110
1 10
120
55
330
400
185
Rotors
1
2
2
2
2
1
3
4
2
Compression
Ratio
8.6:1
9.0:1
9.4:1
9.4:1
9.4:1
9. 0:1
9.3:1
9.3:1
8.5:1
-------
Table 7-2. Rotary Engine Developers
Manufacturer
Approximate HP Range
Area of Application
Ul
Curtis s-Wright
Mercedes Benz
Fichtel and Sachs
Graupner
NSU
Outboard Marine
Rolls-Royce
Toyo Kogyo
Yanmar
General Motors
up to 780
up to 350
3 to 23
0.63
3-136
35
55-180
90-190
22-50
Automotive, truck, aircraft,
marine, power generation,
outboard motors, etc.
Automotive
Lawnmowers, minibikes,
snowmobiles
Model airplanes
Automotive, portable pumps,
marine
Marine, snowmobile
Two-stage diesel cycle
Automotive
Outboard motors
Automotive (1975 model year
Chevrolet Vega)
-------
COOLING FAN
WATER PUMP
END
COVER
ROTOR HOUSING
ROTORS
HYDRAULIC TORQUE-
CONVERTER
CLUTCH
OIL PUMP
OIL FILTER
MAINSHAFT
SEPARATING WALLS
Figure 7-2. Production Rotary Piston Engine for Automotive
Application, Cutaway Drawing (Ref. 7-1)
-------
(A)
INTAKE EXHAUST
INTAKE EXHAUST
(C)
(D)
t ' /
INTAKE EXHAUST
INTAKE EXHAUST
Figure 7-3. Motion of Rotor in Main Housing of Rotary
Piston Engine (Ref. 7-2)
77
-------
shape and location of the ports and by the relative position of the rotor.
As the rotor apex passes over each port, it transfers the connection of that
port from the chamber leading the apex to the chamber trailing the apex.
The motion of the rotor in its orbit is determined by the
rolling action of the rotor internal gear (large circle in Figure 7-3) about
the central stationary gear (small circle) which is flexed to one of the end
plates. This motion of the rotor transmits power to the mainshaft eccentric
(also represented by the large circle in Figure 7-3) which in turn causes
the mainshaft (small circle) to rotate within the stationary gear and main
bearing.
The Wankel geometry is such that the ratio of mainshaft-to-
rotor rotation is 3:1. Since each rotor flank completes one power stroke
per rotor revolution, the rotor completes three power strokes per rotor
revolution or one power stroke per mainshaft revolution. On the basis of
power strokes per revolution, a one-rotor engine would be equivalent to a
single cylinder, two-cycle reciprocating engine or a two-cylinder, four-
cycle reciprocating engine.
An important design parameter of the rotary-piston engine
is called the K-factor, and its value determines the shape of the inner sur-
face of the rotor housing, the shape of the rotor, the size of the engine for
a given swept volume, and the maximum compression ratio. The factor is
defined by the following ratio:
where
R = radius to the rotor apex
E = radius of the mainshaft eccentricity
Figure 7-4 shows the variation in relative size of an engine having a fixed
swept volume and different values for the K-factor. Also shown are the
maximum theoretical compression ratios for each engine.
7-8
-------
~T = 17-5
C.R. = 30
C.R. = 18
-T = 3-9
C.R. = 10
Figure 7-4. Variation of Size and Compression Ratio for
Engines with Fixed Swept Volume and
Various K-Factors (Ref. 7-1)
7-9
-------
The following general trends are indicated when the K-factor
is increased and the swept volume is held constant:
• The shape of the rotor housing becomes less "pinched"
at the waist
• The overall size of the engine is increased
• The maximum compression ratio is increased.
Current automotive rotary-piston engines are designed with
a K-factor of about 7.
1.2.2 Design Features
One of the highly desirable characteristics of this engine is
its considerable reduction in basic complexity, as compared with the re-
ciprocating engine. Elimination of the complete valve train and connecting
rods results in lowering the mechanical noise level, greatly simplifies the
balancing, and is an area of cost reduction. Because all major masses
revolve about a fixed center of rotation, balancing can be accomplished by
the simple use of counter weights. The net effect is to provide an engine
with smoother and quieter operation than the reciprocating piston engine
and with higher peak speed capability. Figure 7-5 compares the rotating
parts of the rotary engine with that of a two-cylinder piston engine to illus-
trate the difference in complexity.
For a given power rating, the rotary engine is lighter and smaller
than a reciprocating piston spark ignition engine. Values of weight-to-power
ratio range from 2.0 to 2.4 Ib/hp for current production versions of rotary auto-
motive engines. This can be compared to a typical value of about 3. 6 Ib/hp
for a spark ignition reciprocating engine. One manufacturer indicates red-
uctions of 30 percent in engine weight and 50 percent in size from a compar-
able piston engine. Reductions in engine size and weight can also lead to
smaller-size, lighter-weight vehicles as a result of more compact packag-
ing. The size effect is illustrated by a profile comparison with a V-8 piston
7-10
-------
Figure 7-5. Comparison of Rotating Parts (Ref. 7-3)
-------
engine in Figure 7-6. Figure 7-7 compares the number of parts used in
the Mazda RX-2 rotary engine and the main parts of a six-cylinder, 2.7-
liter reciprocating engine.
Mechanical design flexibility is a characteristic of this
engine. Design variations include single and multirotors, liquid cooling and
air cooling, spark and compression ignition, homogeneous and stratified
charge, carbureted and injected engines, single and dual spark plugs,
peripheral and side ports, single and multiports, and extended speed range.
This flexibility provides opportunity for optimizing the mechanical design
of this engine with regard to combustion chamber shape, spark plug location
and plug-port designs, and the use of multiports for stratified charge
applications.
Fuel, lubrication, filtration, and cooling requirements are
very similar to those for the spark-ignition reciprocating engine. However,
there are some unique cooling and material requirements related to apex
seals and engine thermal gradients. High thermal stress in this engine
results from the fact that different parts of the engine are exposed to con-
ditions of a near-constant thermal environment, but the environment varies
drastically from one part of the engine to another. The intake side of the
engine is continually exposed to only cool gases, while the exhaust side sees
only hot exhaust gases. The housing in the vicinity of the spark plug is
exposed almost continuously to a very hot flame, giving rise to excessively
high thermal gradients in this region. The only reasonable solution to
relieve thermal gradients is to use a housing material, such as aluminum,
with good heat transfer characteristics to quickly dissipate the heat and
thereby reduce temperature gradients. The use of cast iron in this region
has not been successful to date because of its low thermal conductivity and
reduced resistance to tensile stresses brought about by high thermal
gradients incurred by engine transients.
7-12
-------
I
I—"
OO
V-8
Figure 7-6. Silhouette Comparison of a Twin-Rotor
Wankel Engine (Curtiss-Wright
RC 2-60 U5) with a 283-cubic inch
Chevrolet V-8 Piston Engine
(Ref. 7-1)
-------
COCO C3CX3O
a
J6. i.
'. \ \
( t
t *
t I
* I
s>^
o
i =» i
i - - -
;4 » 1 <
Figure 7-7. Comparison of Main Parts for RX-Z Rotary Engine and 6-Cylinder
Reciprocating Engine (Ref. 7- 4)
-------
7.2.3 Operating Characteristics
Engine operating characteristics are essentially identical to
those for the spark-ignition, reciprocating piston engine, and the rotary
engine can be linked to conventional transmissions. Transmission-gear
ratios may be revised somewhat to compensate for the fact that the rotary
engine achieves high torque at higher peak speeds, similar in this charac-
teristic to some small, four-cylinder reciprocating engines.
7.3 PERFORMANCE CHARACTERISTICS
The basic engine produces high levels of hydrocarbon and
carbon monoxide exhaust emissions. These result from seal leakage, wall
quenching from high surface-to-volume ratios and crevices, and incomplete
combustion (Ref. 7-5). The same factors contribute to low combustion
temperatures and resulting low levels of NO .
The control of HC and CO emissions in current (1973) produc-
tion rotary engines is accomplished by the use of air injection into a thermal
reactor coupled to the exhaust ports. Typical results achieved on the 1973
model year Mazda (RX-3) are shown in Table 7-3, together with results of
development tests performed by Toyo Kogyo and General Motors with the
objective of meeting the original 1975 standards with the thermal reactor
system (Ref. 7-6).
The best results achieved with oxidation catalysts are shown
in Table 7-4. Results at low mileage are seen to approach the 1976 stan-
dards but do not meet the 1977 NO standards (Ref. 7-6). It should be
noted that General Motors (Ref. 7-7), Ford (Ref. 7-8), and Daimler-Benz
(Ref. 7-9) have all reported extremely poor durability results to date for
oxidation catalyst systems used with rotary engines.
The fuel economy of 1973 model year rotary engine cars
(i.e., Mazda) is approximately 30 percent below piston engine vehicles of
the same inertia weight (Ref. 7-10). Development work reported by General
7-15
-------
Table 7-3. Best Emissions Results, Thermal Reactor and Wankel Engine,
2,750-lb Compact Car (Ref. 7-6)
-j
i
Manufacturer
Toyo Kogyo
(Mazda)
Toyo Kogyo
General
Motors
System*
Reactor
Reactor
Reactor 4
EGR
Reactor
(best effort)
Mileage
4, 000
50, 000
Low
Low
Low
No. of Tests
1973
Certification
Many
3
8
72
1
Emissions in
gm/mi ##
HC
2. 4
0. 32
0. 36
0. 35
0. 60
0.43
CO
20. 0
3. 1
2. 6
2. 2
5. 0
2. 8
NOX
0. 9
0. 83
0. 87
0.49
0. 60
0.44
Fuel
r^ena ity
percent
0
(Baseline)
6. 5
--
12. 0
--
--
*A11 these systems are carbureted fuel rich (about 12:1)
**1975 CVS-CH test procedure
##*Relative to 1973 production rotary engine car
-------
Table 7-4. Emissions at Low-Mileage Rotary Engine with
Oxidation Catalyst * (Ref. 7-6)
System
Oxidation catalyst and
air pump
(Best Effort)
No. of Tests
22
1
**
Emissions, grams /mile
HC
0. 7
0.4
CO
0.4
0. 2
NOX
1. 2
1. 0
^General Motors data, 2750-lb car
**1975 CVS-CH test procedure
Motors (Ref. 7-7), on a modified Mazda engine (lean carburetor, modified
ignition timing) tailored for fuel economy without emission control devices,
compared favorably to a 1971 Opel 1.9-liter vehicle at 24 to 29 mpg.
Leakage of gases from the combustion chamber is one of
the more perplexing problems introduced by the rotary engine. This leakage
results primarily from the necessary use of an apex seal having a single
line of contact with the outer housing; it has a strong effect on both fuel
economy and exhaust emissions. A view of a typical apex seal is given in
Figure 7-8.
Current apex seal materials appear to have a reasonable
wear life of about 80,000 miles. However, even with new seals, gas leakage
is excessive. This appears to be due in part to the hydrodynamic effect
resulting in seal lift from the housing surface and the subsequent leakage of
high pressure combustion gases into both the leading and trailing chambers.
Gases leaking into the leading chamber have a high concentration of unburned
7-17
-------
APEX SEAL
RETAINER PIN
SPRING
SIDE SEAL
SPRING
SPRING
Figure 7-8. Typical Apex Seal (Ref. 7-1)
mixture, leading to loss of power (poor fuel economy) and to high hydro-
carbon concentrations in the exhaust. A lowering of combustion chamber
peak pressure also results.
Improvements in sealing will reduce the mixture loss to the
leading chamber, thus improving fuel economy and decreasing hydrocarbon
emissions; but this change will also increase peak combustion pressures,
thereby increasing the production of NO . Therefore, it can be expected
that improved sealing will tend to correct the fuel economy and hydrocarbon
emission problems, but will contribute to the NO emission problem.
7.4
CURRENT STATUS OF TECHNOLOGY
Rotary engine-powered cars constitute a very small fraction
of the current automobile population, but production levels have been rising.
Toyo Kogyo is presently the primary producer of these engines, but they
may be surpassed by General Motors if the Chevrolet Vega rotary engine is
produced in model year 1975, as scheduled, and finds large public
acceptance.
7-18
-------
Available information indicates a lower rate of production
(engines/hour/line) as compared with the piston engine. This results from
the complex machining operations that are required to grind and finish the
working surfaces of both the housing and rotor. Published information
(Ref. 7-11) on production machines give capabilities of around 25 housings
per hour per line as compared with typical production rates for piston
engines of over 100 per hour. This limitation appears to have a significant
impact on manufacturing costs. Either a major breakthrough in the manu-
facturing process or the use of parallel production lines will be required to
achieve current piston-engine production rates.
7.5 PROJECTED STATUS
The more important problem areas facing the rotary-engine
designer are:
• Gas seal leakage
• Low speed fuel economy
• High level of exhaust emissions
• High thermal stresses.
Low fuel economy^ especially at low speeds, is related to
the leakage problem and to the characteristic slow burning rate of the
rotary engine. A solution to the sealing problem will help improve fuel
economy. The slow burning rate, as evidenced in pressure-volume plots,
shows a drastic departure from the ideal constant-volume combustion of
the Otto Cycle, resulting in a drop in efficiency. The use of dual ignition
ports could produce some increase in the burning rate, which may improve
combustion characteristics.
The high surface-to-volume ratio is an inherent characteristic of
this engine and cannot be modified to any great extent. Crevice effects can
be reduced down to some minimum by proper design, but they cannot be
eliminated. Seal leakage has already been discussed and potentially large
improvements can be made here, while incomplete combustion can be
improved with further research leading to a better understanding of the
7-19
-------
unique combustion process in the rotary engine. Stratified charge
combustion is a strong possibility in this engine, which appears to offer
lower emissions and higher fuel economy (Ref. 7-12).
Further optimization in rotary engine design, including the use
of stratified charge combustion and improved versions of apex seals and
thermal reactors, may overcome current fuel economy and emissions
problems of the basic engine. (With the addition of a thermal reactor, the
Mazda rotary engine has demonstrated the capability of meeting the original
1975 Federal emission standards. )
7-20
-------
SECTION 8
-------
8. STRATIFIED CHARGE ENGINES
8.1 INTRODUCTION
8.1.1 General Description
In principle, the stratified charge engine represents a modifi-
cation of the conventional spark ignition engine. The main objective of the
modification is to achieve heterogeneous combustion in which a rich fuel
mixture is generated around the spark plug and a lean mixture is generated
in remaining combustion chamber zones. The resulting two-stage com-
bustion process permits operation of the engine at very lean overall fuel-air
ratios which is conducive to low emissions, good fuel economy, and reduced
sensitivity of the engine to fuel octane and cetane numbers.
Stratified charge engines can be divided into two distinct
classes: open chamber and divided chamber. In open-chamber configur-
ations, exemplified by the Texaco TCCS and Ford PROCO engines, a single
combustion chamber similar to that of conventional spark ignition engines
is used. During engine operation, an air swirl is set up in the cylinders by
means of directional intake porting combined with special piston cup designs.
Fuel is injected into each cylinder toward the end of the compression stroke.
Upon ignition of the swirling rich mixture surrounding the spark plug, the
burning charge expands into the combustion chamber's outer regions where
the combustion process is completed in an oxygen-rich environment. Att-
empts are currently being made to replace the more expensive fuel injec-
tion systems employed in these engines with conventional carburetors.
The divided-chamber stratified charge engines or prechamber
engines, exemplified by Honda's CVCC engine concept, employ two intercon-
nected combustion chambers per cylinder. During the induction stroke of the
piston, a fuel-rich mixture is inducted into the generally smaller prechamber
while the main chamber is charged with a lean mixture or even pure air.
8-1
-------
Upon ignition in the prechamber, hot gases expand into the main chamber
where combustion is then carried to completion. The principal advantage
of prechamber engines over conventional engines is their ability to operate
with very lean overall fuel air mixtures resulting in low emissions, par-
ticularly NO . However, because of the less favorable combustion chamber
surface-to-volume ratio combined with high turbulence, the heat losses of this
engine tend to be higher than in conventional designs. The benefits in terms of
emission reduction and fuel economy improvement that might be realized
in a particular design depend upon the tradeoffs between the heat losses and
the inherently higher thermodynamic cycle efficiency obtained with operation
in the lean fuel air mixture regime.
8.1.2 Historical Developments
The concept of stratified charge operation of internal combustion
engines and its potential was first evaluated in the early 1920s by Ricardo in
England. In the United States, work on the open-chamber configurations
commenced in the 1940s with Texaco's invention of a "knockless" engine.
Research and development of a number of alternate designs have since been
pursued by several investigators and organizations throughout the world. In
1965, the United States Army Tank Automotive Command (TACOM) initiated
funding of exploratory research of both open-chamber and divided-chamber
stratified charge engine concepts which culminated in the development efforts
conducted over a number of years by Ford and Texaco. The objective of the
TACOM-sponsored Ford and Texaco programs was the development of a
military engine with multifuel capability and improved fuel economy relative
to conventional spark ignition engines. Minimization of exhaust emissions
was added later as a design requirement under joint EPA and TACOM
sponsorship. The level of funding of the Texaco effort had been recently in-
creased prior to completion of the program. The TACOM support of the
Ford development program was previously terminated. However, Ford is
conducting a number of in-house programs aimed at the development of its
stratitied charge engine concepts for potential use in passenger cars.
8-2
-------
The concept of a divided combustion chamber can be traced
back to the first "oil engines" in the pre-Diesel era. These early pre-
chamber engines operated on naphtha or heavy oils that were injected into a
spherical prechamber and vaporized in contact with the red-hot thermally
insulated prechamber walls. Later, attempts were made to apply the pre-
chamber concept to the spark ignition gasoline engine, as documented by
numerous patents. However, to date, a practical and widespread application
of the prechamber concept has been realized only in high-speed diesel engines,
principally due to the pioneering efforts of Ricardo. Currently, a number of
organizations and individuals are working on the development of prechamber
engines. In particular, the Honda Motor Company of Japan has conducted
extensive design, development, and test work on its CVCC concept during
the past several years.
8-2 OPEN-CHAMBER STRATIFIED CHARGE ENGINES
8.2.1 General
Since, as previously noted, Ricardo developed the engineering
concept of the stratified charge operation (early 1920s), and since Texaco's
invention of the "knockless engine" in the 1940s, many investigators have
conducted research on a number of alternate designs throughout the world.
In 1965 the U.S. Army Tank Automotive Command (USATACOM) started
to fund exploratory research on various stratified charge engine concepts in
an effort to develop an engine with multifuel capability and better fuel economy
than conventional spark ignition engines. These engine designs included the
Texaco Combustion Process, the Borg Warner Combustion System, the
Witzky Swirl Stratification System, and two divided chamber configurations
then under development at Continental Aviation and Engineering and at the
University of Rochester. Based on this initial work, TACOM subsequently
selected the Ford Combustion Process (FCP) and the Texaco Combustion
Process (TCP) for further development and evaluation (Ref. 8-1). Low
exhaust emissions were added later as a design requirement under joint
TACOM and EPA funding.
5-3
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8.2.2 Power Plant Description
A number of the open chamber, stratified charge engine
concepts developed to date are described in the following sections. These
include the Baudry engine, the Borg Warner Combustion System, the
Hesselmann engine, the Mitsubishi Combustion Process, the Witzky engine,
the Texaco Controlled Combustion System, and the Ford Combustion Process.
8.2.2.1 Baudry Engine
This engine concept, illustrated in Figure 8-1, was developed
by the French Petroleum Institute (Ref. 8-2). As shown in the figure, a
small tube provides a fuel-rich mixture at the cylinder inlet. This mixture
is then carried by the in-rushing air towards the spark plug which is located
at the periphery of the combustion chamber. Although the operation of the
engine is difficult to explain, Baudry reports remarkable improvements in
part-load fuel economy and CO emissions relative to conventional spark
ignition engines, particularly in the low power regime.
8.2.2.2 Borg Warner Combustion System
The Borg Warner Combustion System is shown in Figure 8-2.
In this concept, charge stratification is achieved by injecting the fuel into
the piston cavity at bottom dead center. The fuel-rich combustion mixture
is contained in the cavity during the compression stroke, and combustion is
initiated by means of a spark plug near piston top dead center (Ref. 8-1).
8.2.2.3 Hesselmann Engine
The Hesselmann engine, shown schematically in Figure 8-3, was
designed in the early 1930s with the intention to combine the advantages of
the Diesel and Otto cycle engines. This engine, which is also known as the
Hesselmann "low-pressure diesel," utilizes a high air swirl in the combus-
tion chamber, generated by shrouded intake valves. The fuel is injected
towards the center of the cylinder through a nozzle positioned at the periphery
8-4
-------
co
RICH MIXTURE TUBE
Figure 8-1. Baudry Stratified Charge Engine (Ref. 8-2)
-------
FUEL INJECTOR
VALVE
SPARK PLUG-
COMBUSTION
CHAMBER ^
PISTON POSITION
AT TOP
DEAD CENTER
PISTON POSITION
AT BOTTOM
DEAD CENTER
Figure 8-2. Borg Warner Combustion System (Ref. 8-1)
8-6
-------
Figure 8-3. Hesselmann Stratified Charge Engine (Ref. 8-3)
of the combustion chamber. Fuel injection occurs during the compression
stroke of the piston. Ignition of the combustible fuel- air mixture is
accomplished after completion of the injection process by means of a
spark plug located diametrically opposite the injection nozzle. Except
at idle, the engine operates unthrottled, and its power output is controlled
by the fuel supply (Ref. 8-3).
8.2.2.4
Mitsubishi Combustion Process
The Mitsubishi Combustion Process (MCP) is illustrated in
Figure 8-4 (Ref. 8-4). In this engine, an air swirl is generated by spiral
induction ports, and the fuel is injected against this swirl into the piston cav-
ity. Tests conducted by Mitsubishi indicate low emission and fuel consump-
tion characteristics, especially at light loads. Also, the engine has been
run successfully on a variety of fuels including gasoline, kerosene, fuel oil,
and diesel fuel. The engine exhibits good starting characteristics, even
at temperatures as low as -22°F, and its noise and vibration levels are
comparable to conventional spark ignition engines (Ref. 8-4).
-------
SPARK
PLUG
INJECTION
NOZZLE
Figure 8-4. Mitsubishi Combustion Process (Ref. 8-4)
8.2.2.5 Witzky Engine
This engine concept is depicted in Figure 8-5. It is based on the
unthrottled Otto cycle process and utilizes direct fuel injection into the cylinder
during the last portion of the compression stroke. The fuel is injected against
an air swirl, set up by special intake ports, in such a manner that a fuel-rich
mixture is formed in the vicinity of a centrally located spark plug (Ref 8-5).
Upon ignition in the rich zone, the flame spreads into the lean outer regions
of the combustion chamber. The power output of the engine is controlled
by varying the amount of fuel injected. According to Witzky, the engine
has been successfully operated at very lean air-fuel mixtures and with low
octane fuels. Other advantages are claimed to include smooth combustion
with low peak pressures and pressure rise rates, good part-load fuel
economy, low combustion noise, good acceleration characteristics, and
ease of starting even at low ambient temperatures (Ref. 8-3).
1-8
-------
INTAKE VALVE
SPARK PLUG
INTAKE SWIRL PORT
INJECTION NOZZLE
Figure 8-5. Witzky Stratified Charge Engine (Ref. 8-5)
The Witzky system has been incorporated by the Southwest
Research Institute into a 215-CID Buick aluminum V-8 engine (Ref. 8-3). In
this installation, every other cylinder in the firing order was converted to
stratified charge operation, while the remaining four cylinders remained
unmodified. Figure 8-6 shows the measured full-throttle performance of
the "four-cylinder" stratified charge engine adjusted to account for the
friction losses of the four nonoperative cylinders. Also shown in this
figure are performance curves for the "four-cylinder" carbureted baseline
engine and for the Mercedes Benz 190D light-duty automobile diesel engine.
As indicated, the specific fuel consumption of the Witzky stratified charge
engine is comparable to that of the diesel and substantially below the
baseline engine, particularly at low speeds.
8-9
-------
•STRATIFIED
•CARBURETED
DISPLACEMENT 107.5 in.
CYLINEE
DISPLACEMENT 107.5 in.3
CYLINDERS 4
93 OCTANE
100 OCTANE
3
DIESEL
DISPLACEMENT 115 in.'
CYLINDERS 4
1000 2000 3000 4000
ENGINE SPEED, rpm
Figure 8-6. Full Throttle Performance of Witzky Stratified
Charge Engine, Conventional Gasoline Engine,
and Diesel Engine (Ref. 8-3)
8.2.2.6 Texaco Controlled Combustion System
8.2.2.6. 1 Concept Description
The Texaco Controlled Combustion System (TCCS), illustrated
in Figures 8-7 and 8-8, is a stratified charge process utilizing controlled air
swirl and fuel injection combined with positive spark ignition (Refs. 8-6
through 8-8). Injection of the fuel into the swirling air commences near
the end of the compression stroke, and the injected fuel is immediately
ignited to establish a flame front adjacent to the injection nozzle. As injec-
tion continues, additional fuel is supplied to the flame front and burned very
rapidly as soon as a combustible mixture has been formed.
8-10
-------
DIRECTION OF
AIR SWIRL
NOZZLE
SPARK
PLUG
1. FUEL SPRAY
2. FUEL - AIR MIXING ZONE
3. FLAME FRONT AREA
4. COMBUSTION PRODUCTS
Figure 8-7- Texaco Controlled Combustion System
5-11
-------
SPARK PLUG
Figure 8-8. Texaco Cup Combustion Chamber,
Stratified Charge Engine (Ref. 8-8)
8-12
-------
The TCCS engine is designed to operate unthrottled with the
power output being controlled by the duration and amount of fuel injected into
the combustion chamber. At full load, the injection duration covers about
one revolution of the air swirl. Lean fuel air mixtures are utilized at part-
load and near-stoichiometric mixtures are used at full load. Because of the
lean mixture operation, the engine exhibits excellent part-load efficiencies.
However, at full load, the fuel air mixing process in the TCCS engine is not
as efficient as in conventional spark ignition engines and, as a result, the
maximum power output of TCCS is smoke limited. Owing to the direct cylin-
der injection and positive ignition, the engine has excellent throttle response
and very good starting and warmup characteristics.
Since the TCCS engine utilizes controlled air swirl, fuel injec-
tion, and positive ignition, it is not penalized by some of the inherent disad-
vantages of conventional spark ignition and diesel engines. Ignition of the
first increment of fuel at the time of injection eliminates the possibility of
mixture preignition and engine knock frequently encountered in gasoline
engines, regardless of the octane number of the fuel used. Relative to diesels,
the positive ignition utilized in the TCCS engine minimizes the ignition delay
and the sensitivity of the engine to fuel cetane number. Since the ignition
delay is very short, the peak cylinder pressure and the pressure rise rates
are considerably lower than in diesel engines. Therefore, the mechanical
loads and shocks occurring during combustion are less severe, permitting
the use of lightweight engine blocks, cylinder heads, crankshafts, and pistons.
Work on the TCCS concept has been under way at Texaco since
the early 1940s. Many single-cylinder and multicylinder configurations have
since been evaluated for potential application in industrial and automotive
installations. Since 1965, Texaco has been under contract to TACOM to
adapt the TCCS concept to the existing military light-duty L-141 engine which
is used in the M-151 jeep vehicle. The principal design objectives established
by TACOM include good fuel economy, broad fuel tolerance, and good
driveability.
8-13
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Both naturally-aspirated and turbocharged TCCS engines have
been evaluated by Texaco. The design features of these two concepts are
described in the following sections.
8.2.2.6.2 Naturally Aspirated L-141 TCCS Engine
The L-141 engine is a water-cooled, four-cylinder, in-line spark
ignition engine having a cylinder bore of 3.875 inches, a stroke of 3.00 inches,
and a total displacement of 141.5 cubic inches. The specifications of the
standard L-141 engine and the TCCS modification are listed in Table 8-1.
Because of the oversquare bore-to-stroke ratio of the engine,
incorporation of a cup-in-piston combustion chamber is required to ensure
adequate swirl rates during combustion. As shown in Figure 8-8, the piston
cup has a cylindrical upper portion and a toroidal bottom. The cup diameter
of about one-half of the bore has been determined by Texaco to be the optimum
for this engine with respect to swirl generation and combustion efficiency.
The remaining parts of the piston assembly are standard L-141 items. In
the selected chamber arrangement, the swirl increases substantially near
the end of the compression stroke by transferring the rotating air charge
into the small cup. As a result, a rather low intake swirl rate can be
applied, which is desirable from a volumetric efficiency point of view.
The cylinder head and the fuel supply system used in the TCCS
engine represent the two components which require the greatest modification
from conventional spark ignition engine design practice. These modifications
involve selection of the proper location and shape of the fuel injection nozzle,
spark plug, and intake port design. Compared with a standard spark ignition
engine, the cylinder head of the TCCS engine is rather simple, having a flat
inner surface with valves located perpendicular to this surface. The fuel
injection pump of the TCCS engine is a commercially available high-pressure
diesel unit. Both the pump and the ignition distributor are driven by a
specially designed drive mechanism.
8-14
-------
Table 8-1. L-141 Engine Specifications, Standard and TCCS
(Refs. 8-7, 8-8)
Standard
TCCS
. 3
Configur at ion
Cycle
Cooling
Number of Cylinders
Total Displacement, in
Bore, in.
Stroke, in.
Compression Ratio
Fuel
Induction System
Firing Order
Weight (dry with accessories), Ib
Length, in.
Width, in.
Height, in.
Maximum Rated Power, Bhp
Rated Speed, rpm
Maximum Torque, Ib-ft
Speed at Maximum Torque, rpm
7. 5
Combat Gas
Naturally
Aspirated
1-3-4-2
328
26.9
21.4
23.0
65
4000
110
1800
(a)
In-line
4-stroke
Liquid
Four
141.5
3.875
3.00
10.0
Gas through
Diesel No. 2
Naturally
Aspirated
1-3-4-2
375
27. 5
24. 5
23.3
64
3200
2400
*a'At 28.77 in Hg and 85°F, with air cleaner, generator, and automatic
spark advance
'At 29. 92 in Hg and 60°F, with air cleaner, generator, and automatic spark
and injection advance
8-15
-------
An emission-control system, consisting of EGR, oxidation
catalysts, and intake-air throttling, was added later in the program at the
expense of some loss in fuel economy. This system is further described in
Section 8.2.7-2.2.
With the exception of these components, the design and
construction of the TCCS engine is comparable to other reciprocating
engines. The TCCS engine has been operated up to the maximum governed
speed of 3800 rpm. Texaco feels that speeds up to 5000 rpm are feasible with
TCCS by modifying the intake system and by utilizing a more favorable
bore-to-stroke ratio.
8.2.2.6.3 Turbocharged L-141 TCCS Engine
Following the successful application of engine intake pressure
charging in exploratory tests, Texaco has incorporated a commercially
available turbocharger into some of its L-141 TCCS engines. Other changes
to the engine were minor, with the most significant being a reduction of the
compression ratio from 10:1 to 9.3:1, to reduce the structural loading of the
lightweight L-141 engine and to minimize engine knock at full power
(Ref. 8-8).
On the dynamometer, the engine has demonstrated good
operating characteristics and is free from misfire and knock when operated
on combat gasoline, CITE fuel, and No. 2 Diesel fuel. The effects of fuel
variations on engine performance are further discussed in Section 8. 2.7. 1.
8.2.2.7 Ford Programmed Combustion Concept
8.2.2.7. 1 General
Ford's Programmed Combustion (PROCO) concept represents
a second-generation modification to the current reciprocating internal com-
bustion engine. The principal design objectives of the first-generation Ford
Combustion Process (FCP) engine were improved fuel economy compared
with conventional spark ignition engines (without power loss), adequate cold
8-16
-------
Itart characteristics, a high degree of parts commonality with carbureted
engines; smooth operating characteristics; and low emissions (Refs. 8-9
and 8-10). Research and development on this concept was initiated by Ford
in 1956, and, starting in 1968, the design was applied to the L-141 military
light-duty engine under TACOM sponsorship. Although the emissions of the
original FCP engine were respectable, they fell far short of meeting long-
range objectives.
Development of the second-generation PROCO engine was
initiated in 1969 under joint TACOM and EPA sponsorship. In addition to
meeting the objectives of the FCP engine, PROCO was designed to meet the
1976 statutory federal emission standards at low mileage with a minimum
loss in fuel economy. These objectives have been achieved.
TACOM terminated its support of the L-141 PROCO program
as of March 15, 19*73 and is concentrating its efforts on development of the
Texaco concept. However, work on PROCO and on carbureted modifications
of the PROCO concept (Fast Burn) is being continued by Ford with company
funding (Ref. 8-11). These efforts are aimed at the development of PROCO
and Fast Burn engines for use in passenger cars. The design features and
operational characteristics of these engines are outlined in the following
sections.
8.2.2.7.2 PROCO Engine
The PROCO engine concept, as applied to the military L-141
light-duty engine, is illustrated in Figure 8-9. Engine specifications are
listed in Table 8-2. In this engine, air enters through a special intake system
which is designed to impart a moderate swirl around the cylinder bore axis
in order to assist in attaining proper fuel-air mixing and good combustion.
The air charge inducted into the cylinder is then compressed and transferred
into the cup-shaped combustion chamber located concentrically in the piston.
The top opening of the chamber has been optimized and represents 35 percent
of the bore area, resulting in 65 percent squish action (Ref. 8-10).
8-17
-------
INJECTOR ASSY
/SPARK PLUG
COMBUSTION
CHAMBER
Figure 8-9. Ford L-141 PROCO Engine (Ref. 8-10)
8-18
-------
Table 8-2. PROCO Specifications, L-141 Engine (Ref. 8-10)
Number of Cylinders
Cycle
Displacement, CID
Compression Ratio
Fuel Injection System
Air-Fuel Ratio Control
Ignition System
Spark Plugs
EGR System
Exhaust Manifold
Catalyst
Fuel
Four-stroke
141.5
11.0
Ford Experimental
Speed - Vacuum
Ford Transistorized
Experimental Short Electrode
Fixed Orifice, water-cooled
Insulated, low inertia
Engelhard PTX-6 (noble metal)
Lead Sterile; 91 RON
8-19
-------
In addition, Ford has incorporated the PROCO concept into
several of its 351-CID V8 engines. With the exception of a symmetrical
piston cup and a dished cylinder head surface, the PROCO components used
in this engine are equivalent to those employed in the L-141 engine
(Ref. 8-10).
A compression ratio of 11.0 has been selected by Ford for
all PROCO engines. Fuel is injected into each cylinder through an inclined
nozzle configured to provide a low, penetrating, wide-angle (approximately
100 degrees), conical fuel spray. The spray mixes with the air and forms a
rich mixture in its center surrounded by increasingly leaner mixtures and
a layer of air adjacent to the cylinder wall to minimize quenching effects.
A specially designed sparkplug located just above the spray cone is used
to ignite the mixture near the top-dead-center (TDC) position of the piston.
From there, combustion proceeds very rapidly into the rich mixture cone and
then into the leaner regions. As the piston descends, the swirling charge
expands from the piston cup into the cylinder where the combustion process
is then carried to completion.
The NO emissions of the engine are further reduced by means
ji
of exhaust gas recirculation (EGR), which is inducted into the air intake after
passing through a water-cooled heat exchanger. The HC and CO emissions
are controlled by a system consisting of a low-inertia, insulated manifold;
an oxidation catalyst, and air throttling.
The HC, CO, and NO emissions from the PROCO engine are
inherently low due to the stratified charge combustion process. HC is
minimized internally through the use of late injection in conjunction with
a soft fuel spray used to prevent fuel deposition on the cylinder wall.
CO is inherently low because of the availability of excess air at all operating
conditions except at full power. NO is controlled by initiating combustion in
the fuel-rich zone, which results in moderate flame temperature and oxygen-
free, post-flame gases. In addition, the engine is operated with high EGR
rates and late combustion to further reduce the combustion temperature and
the NOx formation rates. PROCO engine and vehicle emission performance
and test data are discussed in Sections 8.2.7.2.3 and 8.2.7.3.2.
8-20
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8.2.2.7.3 Fast Burn Concept
The Ford Fast Burn engine concept is a modification of the
PROCO engine in which the fuel injection system has been replaced by a
carburetor. To date, two configurations, Fast Burn Phase I and Fast Burn
Phase II, have been tested by Ford (Ref. 8-11).
Fast Burn Phase I consists of a bowl-in-piston combustion
chamber similar to the PROCO design. The engine has a compression ratio
of 11.0 and utilizes high air swirl rates in conjunction with improved
carburetion, EGR, and modified spark advance. This concept was tested
in a modified 351-CID engine installed in a 1972 Torino automobile. The
emission and performance data from these tests are presented in
Sections 8.2.7.2.3 and 8.2.7.3.2.
The design features of the Fast Burn Phase II concept are
illustrated in Figure 8-10. As indicated, the engine incorporates a
conventional combustion chamber and two spark plugs per cylinder and
utilizes EGR, low-intake air swirl, and a normal production/compression
ratio. Initial concept evaluation tests have been conducted on a 400-CID
V-8 engine installed in a 1973 Torino vehicle.
8.2.3 Fuel Injection and Ignition Systems
Fuel injection in both the TCCS and PROCO engines is aimed
at creating a localized fuel-rich zone in the vicinity of the piston cup. In
the PROCO engine, injection occurs early in the compression stroke so that
rich zone mixing can occur before ignition. To retain combustion stability over
a suitably large speed and load range, the injection timing advance is varied
over a wide range as a function of speed and load. Commercially available
diesel injection systems do not have provisions for such wide timing advance
schedules and they also inject much faster than optimum under the low pres-
sure injection conditions of the PROCO. Therefore, in 1961 Ford initiated
development of a specialized injection pump and nozzles tailored to its
combustion process. The new low-pressure (250 to 350 psi) fuel injection
3-21
-------
DUAL IGNITION
(two spark plugs)
CONVENTIONAL
COMBUSTION
CHAMBER
Figure 8-10. Ford Fast Burn Phase II Stratified Charge
Engine (Ref. 8-11)
8-22
-------
system consists of individual plungers for each cylinder, a metering sleeve
for flow control, and a poppet valve type injector designed to vibrate for
improved atomization.
In the TCCS engine, the fuel is injected at high pressures
provided by a commercially available fuel injection system (Ref. 8-12).
The ignition systems for the two engines also show significant
differences. In PROCO, ignition starts late (essentially at top dead center),
and lasts for only about 3 degrees of crank, permitting the use of a more or
less conventional ignition system. Because fuel injection and spark timing
schedules must be coordinated accurately, the PROCO has a unitized injection
pump/distributor. While this makes for a seemingly complex mechanical
design, the control elements are sealed and protected, and the unit operates
with low fuel pressure and low mechanical loadings.
In the earlier FCP engines, Ford utilized long spark plugs
with electrodes extending to the center of the fuel spray. However, with air
throttling and EGR, shorter electrodes proved to be superior in PROCO with
respect to emissions, fuel consumption, and plug life.
In the Texaco process, sparking is initiated almost simulta-
neously with fuel injection and continues throughout the duration of the injection
cycle. Positive ignition is therefore assured, but a specialized ignition system
is required. The Texaco ignition system features a fast rise time, a 75,000-
volt spark, and multiple sparking (as much as 100 restrikes over a 15 to
20 degree crank angle).
8.2.4 Fuel Requirements
Texaco's naturally aspirated and turbocharged L-141 TCCS
engines have been operated repeatedly on various fuels ranging from high octane
gasoline to high cetane diesel fuel without incurring engine misfire or knock.
Also, the specific fuel consumption and exhaust emissions have remained
essentially unchanged when fuels are switched (Ref. 8-13).
8-23
-------
Conversely, the PROCO engine is susceptible to combustion
knock. However, test data have indicated that the octane sensitivity of PROCO
is less than in conventional spark ignition engines, and 91 RON gasoline is
acceptable. The relatively low octane requirement of PROCO engines is
attributed to the very lean air-fuel mixture of the end gas and the high EGR
rates utilized in this engine.
Since oxidation catalysts are needed for HC and CO control in
these stratified charge engines, lead-free fuel is required to prevent catalyst
poisoning.
8.2.5 Power Control Requirements
The power control requirements and methods utilized in
stratified charge engines are comparable to those in conventional spark
ignition engines.
In the PROCO engine, the injection function incorporates a
compound control system consisting of a speed/density, air fuel ratio control
at idle and part load; a mechanical enrichment system for maximum power;
and a fuel shutoff valve for deceleration. With increasing power demand, the
air throttle is gradually opened to the wide-open position during the first
two-thirds of the accelerator pedal movement while the EGR valve is
kept fully open. During the last third of the pedal movement, the EGR flow
rate is gradually reduced to zero for the full power setting (Ref. 8-10).
In the most recent TCCS engine installations, power is con-
trolled by the accelerator pedal, which is connected to the injection pump and
the automatic transmission. Conversely, the earlier configurations employed
intake throttling at idle and during deceleration, as -well as modulated EGR.
8.2.6 Starting Characteristics
In general, open-chamber stratified charge engines exhibit
good starting characteristics.
Controlled cold-starting tests conducted on the Texaco TCCS
engine at temperatures as low as -20°F indicate excellent starting response
on gasoline. \\'ith CITE fuel, successful starts are usually achieved within
8-24
-------
10 seconds of engine cranking. However, starting problems were occasionally
encountered with diesel fuel (Ref. 8-7).
Ford's stratified charge engines have always exhibited excellent
starting characteristics. Also, driveaway following a cold start has never been
a problem on these engines. However, after prolonged periods of standing, the
engine may idle rough for several seconds, until all the air is purged from
the fuel injection system (Ref. 8-9).
8.2.7 Performance Characteristics
8.2.7. 1 Power and Efficiency
The power output and torque characteristics of Texaco's naturally
aspirated L-141 TCCS engine at full throttle are presented in Figure 8-11 as a
function of engine speed (Ref. 8-8); also shown in this figure are the perform-
ance curves of the conventional L-141 engine. In the mid-speed regime, the
TCCS engine develops somewhat higher power than the standard engine. Con-
versely, in the low-speed regime and near maximum speed, the TCCS engine
has slightly lower power. The loss in engine power at speeds above 3, 200 rpm is
due primarily to injection limitations imposed by engine smoke. Similar smoke-
limited power characteristics were obtained by Texaco for other fuels.
The superior efficiency capability of the TCCS engine concept
with respect to the standard L-141 spark ignition engine is illustrated in
Figure 8-12, showing the brake specific fuel consumption (BSFC) of the two
engines at 2, 800 rpm as a function of brake mean effective pressure (BMEP).
As indicated, the BSFC of the TCCS engine is essentially independent of the
type of fuel used. For a BMEP of 30 psi, the TCCS engine has a 27 percent
lower BSFC and, at 100 psi, the difference is about 24 percent in favor of
TCCS. Similar trends were obtained by Texaco for other engine speeds
(Ref. 8-8). However; when making such comparisons, consideration must
be given to the fact that the efficiency of the standard L-141 engine is lower
than that of the average Otto cycle engine.
The fuel consumption characteristics of the turbocharged
TCCS engine and of Ford's PROCO engine are comparable to the naturally
aspirated TCCS engine.
8-25
-------
CD
I
CM
60
en
o
ac
uj 40
O
a.
UJ
v>
OL 30
O JU
UJ
*
<
QL
ffi
l\>
o
-*
o
800
ENGINE
STD. ENGINE
120
100
80
60
O
a:
O
CD
1600 2400
ENGINE SPEED - RPM
3200
4000
MAP-29.0"HG
GENERATOR
IAT-85°F DYN. EXHAUST AIR CLEANER
AUTOMATIC SPARK AND INJECTION ADVANCE
Figure 8-11. Full-Load Brake Performance, Texaco Naturally Aspirated
L-141 TCCS Engine, Gasoline (Ref. 8-8)
-------
1.00
0.90
0.80
Q.
XI
UJ
U.
CD
i
LL
0.70
0.60
0.50
0.40
0.30
c
•
I
\
1
\
\
\
\
\
\
\
\
\
X
o
d?
[\
\
^
pf-p
+ (O
2800 RPM
O GA
SOLINE
X CITE FUEL
+ DIESEL FUE
STANDARD ENGI
GA
+
X
SOLINE
'-""
hoo>^
:L
NE
^
sX$~
20
40
60 80
BMEP-psi
100
120
140
29.0" HG, AMBIENT PRESS. 85°F, AMBIENT TEMP. DYN. EXHAUST
AIR CLEANER GENERATOR
AUTOMATIC SPARK & INJECTION ADVANCE
Figure 8-12. Multifuel Brake Performance, Texaco
Naturally Aspirated L-141 TCCS Engine
y (Ref. 8-8)
8-27
-------
8.2.7.2 Emissions
8.2.7.2.1 General
A stratified-charge engine inherently emits less pollutants
(HC, CO, NO ) into its exhaust manifold than a carbureted engine with
.X
comparable power output. The reasons for the lower emissions are as
follows:
a. Hydrocarbons - Unburned hydrocarbon emissions result pri-
marily from the incomplete burning of fuel in contact with the
combustion chamber walls. HC emissions also result from
capture of fuel in various crevices and from insufficient
oxygen in the body of the fuel-air charge. In the SCE/1 the
wall and crevice effects are largely eliminated by concen-
tration of fuel in the rich-combustion zone; the walls are in
effect lined with a layer of air rather than a mixture of fuel
and air. Unburned hydrocarbons exist in relatively high
concentrations in the rich zone, but these are readily
burned by the excess oxygen in the lean stage of the burn.
b. Carbon Monoxide - Emissions of CO result whenever the fuel-
air charge is uniformly deficient in oxygen. As in the case
of HC, the rich stage of the SCE process does result in high
initial CO formation, but this is very effectively oxidized to
CO? at the lean end.
c. Oxides of Nitrogen - Three factors play an important and
integrated role in the formation of oxides of nitrogen; peak
cycle temperature, exposure time at high temperature, and
oxygen availability. In the case of the SCE, the initially
rich combustion limits the peak temperature and assures an
oxygen-free environment behind the flame front. By the time
the flame reaches the lean zone, the temperature is reduced,
so the residence time with both high temperature and oxygen
present is greatly minimized. As a result, the NO formed
during combustion is rather low. These effects are depicted in
Figure 8-13. In the early stages of combustion, the three pollut-
ants are formed in concentrations shown on the left of the figure.
By the end of the combustion cycle, the concentrations corre-
spond to the right of the figure, with the transition analogous
to "tunneling" across the figure. Particularly in the case of
NO , the conditions corresponding to near-stoichiometric
combustion are not permitted to exist long enough for
NO to build up to equilibrium levels.
Stratified-Charge Engine.
8-28
-------
END OF
COMBUSTION
RICH
STOICHIOMETRIC
LEAN
AIR FUEL RATIO
Figure 8-13. Effect of Stratified Charge Combustion on Emissions
8-29
-------
NO control to the 1976 level has been achieved by both
x
PROCO and TCCS engines without reduction catalyst after-treatment. Exhaust
gas recirculation alone has proven sufficiently effective to meet this goal.
Further, the oxygen-rich characteristics of the exhaust make the SCE much
more tolerant of high EGR rates than carbureted engines. Recirculation
rates of 25 percent are standard SCE practice, and rates as high as 50 per-
cent (at idle) have been used without causing misfire. The SCE has suffi-
cient NO control margin to permit EGR cutoff at high loads without exceed-
3C
ing the 1976 level, thereby greatly reducing driveability degradation under
peak-load conditions.
8.2.7.2.2 Texaco TCCS Engine
Initial vehicle emission tests on Texaco's TCCS engine were
conducted by the National Air Pollution Control Administration (NAPCA) in
1968. In 1970, NAPCA tested an M-151 vehicle, which was equipped with a
turbocharged L-141 TCCS engine and was driven on the dynamometer over
the then proposed Federal Driving Cycle (LA 4-S3). The emissions from
this engine, measured in accordance with the CVS technique (3,000-pound
inertia weight), are listed along with the emissions of the standard M-151
vehicle in Table 8-3. Both vehicles were tuned for maximum performance
with no consideration given to emissions. Although the uncontrolled TCCS
vehicle has substantially lower emissions than the standard vehicle, the
levels are significantly above the 1975-76 Federal emission standards. As
shown in Table 8-3, some reduction in the HC and CO emissions was
achieved by means of a commercially available oxidation catalyst, which was
installed downstream of the turbocharger.
In an attempt to meet the 1976 standards, Texaco added an
emission control system to its basic TCCS engine design. This system
consists of the following components:
8-30
-------
A cooled, dual-rate EGR system for NOX control,
incorporating a participate trap, an EGR cooler, and an
EGR mixer. At high loads, the EGR flow rate is about
15 percent of the air charge and at low loads the rate is
somewhat higher.
A closely coupled platinum catalyst incorporating a
turbulence device.
A second platinum catalyst and a copper chromite
catalyst installed downstream of the closely coupled
catalyst.
Intake air throttling at idle and low loads in order to
increase the exhaust temperature and improve the
catalyst conversion efficiency at the expense of some
loss in fuel economy.
Table 8-3. Exhaust Emissions of a Turbocharged
M-151 TCCS Vehicle (Ref. 8-8)
^ * / •
Emissions, gm/mi
HC
CO
NO
X
M-151 TCCS
no catalyst
3.85 - 4. 58
9. 08 - 9. 62
1.74 - 1.92
with catalyst
1.74 - 1.97
2.29 - 2.69
1.81 - 2.23
Standard
M-151
6.4
76.2
3.4
'"LA 4-3S Test Cycle
3,000-lb Inertia Weight
Cold Start
Gasoline Fuel
The M-151 vehicle incorporating this baseline emission con-
trol system was tested by the EPA and by Texaco in 1972, and the test results
are presented in Table 8-4. As indicated, vehicle emissions are below the
original 1976 statutory Federal emission standards. However, it should be
noted that the vehicle was not able to either negotiate all the acceleration
modes in the cycle or achieve the maximum cycle speed of 57 mph.
5-31
-------
Table 8-4. Exhaust Emissions, Controlled Naturally Aspirated
M-151 TCCS Vehicle Equipped with Exhaust Gas
Recirculation System and Oxidation Catalysts
Laboratory
Texaco
Texaco
Texaco
EPA
EPA
EPA
Mass Emissions - gm/mi
HC
0.25
0.28
0.40
0.40
0.33
0.37
CO
1.17
0.62
0.26
0.26
0. 15
0.30
NO
X
0.33
0.32
0.29
0.30
0.31
0.31
1975 CVS C/H Refg> 8_8> 8_H
2,750-lb Inertia
Lead-free Gasoline
Upon completion of a 50,000-mile durability run, the vehicle
was retested by the EPA, and the emissions were lower than the 1976 statu-
tory emission standards. However, major maintenance work was required
on the engine during the durability test, including several ignition system
modifications, replacement of four catalysts, cleaning of the EGR system,
and major valve train maintenance (Ref. 8-15).
The effect of emission control system modifications on the
emissions and fuel economy of the baseline TCCS system incorporated into an
M- 151 vehicle is illustrated in Table 8-5. The modifications, which were imple-
mented in 1972, include fuel injection and ignition retardation, various levels
of modulated EGR, and utilization of oxidation catalysts in the engine exhaust.
As indicated in the table, the CO and NO emissions of both the uncontrolled
naturally aspirated, and turbocharged engines are low, compared with
uncontrolled conventional gasoline engines. Conversely, with respect to HC,
3-32
-------
Table 8-5. Effect of Emission Control System Modifications on M-151 TCCS Vehicle
Emissions and Fuel Economy, 1975 FTP
Test
Lab
Texaco
Texaco
Texaco
Texaco
Texaco
Texaco
Texaco
Texaco
Texaco
EPA
EPA
Case
No.
1
2
3
4
5
6
7
8
9
10
11
Engine Configuration
Max. economy setting, no
catalyst
Max. economy setting, no
catalyst
8-deg combustion retard only,
no catalyst
8-deg combustion retard, low
EGR rate, no catalyst
8-deg combustion retard,
medium EGR rate, two
platinum catalysts
13-deg combustion retard,
high EGR rate, two platinum
catalysts
5-deg combustion retard, no
EGR, two platinum catalysts,
alternate vehicle
Baseline configuration''"'"
Baseline configuration with
reduced EGR
Baseline configuration5'"1"
Baseline configuration with
reduced EGR
Turbo-
charged
Yes
No
Yes
Yes
Yes
Yes
Yes
No
No
No
No
Emissions
gm/mi
HC
3. 13
4.24
3.24
3.60
0.33
0.35
0.30
0.36
0.48
0.37
0.50
CO
7.00
7.28
6.43
6.69
1.05
1.41
1.07
0.61
0.57
0.24
0. 14
NOX
1.46
1.43
1.29
0.84
0.61
0.35
1.40
0.31
0.45
0.31
0.70
Fuel
Economy,
mi/ gal
by wt.
24.3
-
22.4
20.5
19.7
16.2
20.9
16.2
17.6
15.8
21.9
Max. Rear
Wheel hp
at 50 mph
40.0
-
39.3
38.0
36.0
28.7
40.0
-
-
-
-
Refer-
ences
8-13
8-13
8-13
8-13
8-13
8-13
8-13
8-8
8-8
8-8
8-8
*Fuel - Gasoline 91 RON, lead-free, + 2% oil.
-'* ;'<
Exhaust Gas Recirculation (EGR) plus two platinum catalysts
oo
i
Oo
-------
the TCCS engines are comparable to conventional spark ignition engines.
Some reduction in HC and CO can be achieved by retarded spark timing at
the expense of lower fuel economy. As in conventional engines, EGR is quite
effective in reducing NO , and the use of oxidation catalysts results in a sub-
stantial reduction of HC and CO. Similar results were obtained with JP-4
and diesel fuel (Ref. 8-8).
Emission data from a TCCS-powered Plymouth Cricket vehicle
tested in 1973 by Texaco, in accordance with the 1975 Federal Test Procedure
(FTP), are listed in Table 8-6. The emission control system incorporated
in the engine consisted of modulated cooled EGR, two oxidation catalysts, and
air throttling at low loads. As indicated, the average emissions obtained in
these tests are below the statutory 1976 standards (Tests 30, 32, and 35).
While CO is attractively low, Chrysler feels that HC and NO are not low
j£
enough to provide an acceptable allowance for system deterioration due to
mileage accumulation and production tolerances (Ref. 8-16). According to
Chrysler, additional development is necessary to simplify the system design
and to improve the EGR control.
Emission and fuel economy data obtained on this vehicle by
Texaco are also listed in Table 8-6. The data demonstrate that injection and
ignition retardation are quite effective in reducing NO emissions at the
jt
expense of some loss in fuel economy. When adjusted for best economy, the
fuel consumption of the TCCS-powered vehicle is about 20 percent lower
than the average 1973 Cricket certification vehicle.
8.2.7.2.3 Ford PROCO and Fast Burn Engines
Results from M-151 PROCO vehicle tests conducted in 1971 by
Ford (Ref. 8-10) and by the EPA (Ref. 8-17) are listed in Table 8-7. In
these tests, the EGR cutoff and fuel enrichment at high power levels were
deactivated to minimize NO . As a result, the vehicle had insufficient power
X.
to negotiate the two highest acceleration modes of the Federal Driving
Cycle. Ford estimates that a 20 to 25 percent larger engine would develop
sufficient power to follow the cycle with EGR.
5-34
-------
Table 8-6. Exhaust Emissions from a Chrysler Cricket TCCS Vehicle,
2,500-lb Inertia Weight (Ref. 8-16)
Test
No.
Avg. 3
(30, 32
& 35)
44
47
50
53
Timing -
Injection and Ignition
Retarded for low emissions
Retarded for low emissions
Retarded for low emissions
Retarded for low emissions
For best economy
Cata-
lysts
Yes
Yes
Yes
Yes
Yes
EGR
Yes
Yes
No
No
No
*
EBP
Yes
No
Yes
No
No
Air
Throt-
tled
Yes
No
No
No
No
1975 EPA gm/mi
HC
0.36
0. 59
0.61
0.73
1. 07
CO
1.15
0.64
0. 85
1. 18
0.84
NOX
0.38
0. 54
0.99
1.20
1.89
Fuel
Economy
mpg
20. 1
22.6
22.5
23.8
25.3
00
I
Ul
EBP = Exhaust Back Pressure Orifice
-------
Table 8-7. Emissions and Fuel Economy, M-151 PROCO Vehicles, 1975
CVS Test Procedure (References 8-10 and 8-17)
Vehicle
PROCO, no catalyst
PROCO, with catalyst
PROCO, with catalyst
Emissions, '
gm/mi
HC
2.60
0.35
0. 37
CO
13.45
1. 01
0.93
NOX
0. 32
0.35
0. 33
Fuel
Economy,' 1)(4)
mi /gal
21. 7
21.3(2)
Not measured
Number of
Tests
(averaged)
1
4
14
Test
Facility
Ford
Ford
EPA
1972 CVS Test Procedure
PROCO, best fuel
economy, no catalyst
PROCO, no catalyst
PROCO, with catalyst
4. 96
3. 10
0. 54
7. 75
13. 75
1. 18
3. 85
0.33
0. 37
23. 8
21.2
19.6
2
1
4
Ford
Ford
Ford
CVS = constant value sampling.
(1) Computed from the mass emission data (miles per gallon).
(2) 20.4 miles per gallon, measured with a burette.
(3) Baseline (carbureted) vehicle emissions are 4. 55 gm/mi HC, 41.6
gm/mi CO, 4.4 gm/mi NOX.
(4) Baseline (carbureted) vehicle fuel economy 17.2 mi/gal.
As indicated in Table 8-7, the low-mileage emissions achieved
with the derated M-151 PROCO vehicle with catalyst are below the 1976 statu-
tory federal emission standards. Ford emphasizes, however, that this was
accomplished with carefully controlled engine adjustments derived from a
long series of optimization tests (Ref. 8-10). Owing to the high EGR rates
employed, the engine operated close to its misfire limit. Without catalyst but
with EGR, the HC emissions of the PROCO engine are about 40 percent lower
than for the carbureted configuration, and CO is about 70 percent lower. At
the best economy setting, PROCO gives about 10 percent lower HC and NOX
and about 85 percent less CO.
In 1973, Ford conducted AMA durability tests of a 1972
Montego automobile equipped with a 351-CID PROCO engine and two Engel-
hard PTX oxidation catalysts. As shown in Table 8-8, the low mileage
8-36
-------
Table 8-8. 351-CID PROCO Durability -- Vehicle: 1972 Montego,
110T722, 1975 Test Procedure (Ref. 8-11)
At the start
At 26, 000 miles,
after repair and
readjustment
At 26, 000 miles,
after application of
part-load back-
pressure system and
240° thermostat
Catalyst
Mileage
10
26,195
26,400
Emissions, gm/mi
HC
0. 16
0.79
0.47
CO
0.25
1. 16
1.11
NOX
0.32
0.37
0.35
Fuel MX
Economy v1/
mpg
13.9
12.75^)
13.2
Based on carbon balance
* ^Tank mileage for 26,000 miles: 12.7 mpg
emissions are below the statutory 1976 standards. However, after accumu-
lation of 26,000 miles, the HC and CO emissions had increased significantly,
indicating excessive catalyst performance degradation. Subsequent installa-
tion of a 240°F thermostat, combined with higher backpressure operation of
the engine at part load, resulted in a substantial reduction in HC (Ref. 8-11).
The effect of EGR on the emissions and fuel economy of this
vehicle, tested without catalysts, is illustrated in Table 8-9. A reduction of
the EGR flow rate from 25 percent to 15 percent resulted in substantially
lower HC emissions at the expense of higher NOx< Apparently, the variations
in EGR performed by Ford had little effect on fuel economy.
Table 8-10 presents emission test data obtained by Ford from
a 351-CID V-8 Fast Burn Phase I engine installed in a 1972 Torino vehicle
(Ref. 8-11). With oxidation catalyst, the vehicle meets the 1976 statutory
emission standards for CO and NO at low mileage and exceeds the HC
!X
standard slightly. However, after 25,000 miles, the HC and CO emissions
8-37
-------
Table 8-9. 351-CID PROCO Emissions and Fuel Economy -- Vehicle:
1972 Montego, 110T722, 1975 Test Procedure (Ref. 8-11)
NOX
Design
Level
(gr/mi)
0.4
1.5
EGR
%
25
15
Emissions, gm/mi
HC
3.54
1.34
CO
28.2
25.0
NOX
.34
.96
Fuel
Economy
mpg
13.8
13.2
Remarks
no catalyst
no catalyst
Based on carbon balance
Table 8-10. Vehicle Emissions and Fuel Economy, 351-CID Fast Burn
Phase I Engine in a 1972 Torino, 1975 Test Procedure,
Average of Two Tests (Ref. 8-11)
Mileage
Low
Low
25, 154
Emissions, gm/mi
HC
6.9
.44
2.52
CO
78. 3
1.31
5.78
NO
X
. 37
.34
.37
Fuel
Economy
mpg
11.9
11.9
11.4
Remarks
w/o catalyst
w/HC-CO catalyst
w/HC-CO catalyst
8-38
-------
are substantially higher than the 1976 standards. The rapid deterioration
of the catalyst is attributed to the high HC and CO levels of the raw engine
exhaust, resulting in high thermal loading of the catalyst.
Fast Burn Phase II vehicle emission data are presented in
Table 8-11 for operation without catalyst and secondary air, but with EGR.
Under these conditions, NO is rather low but HC and CO are excessive.
,x
The data show again the previously noted interrelationship between NO ,
HC, CO, and fuel economy.
Table 8-11. Vehicle Emissions and Fuel Economy, 400-
CID Fast Burn Phase II Engine in a 1973
Torino, 1975 Test Procedure (Ref. 8-11)
NOX
Design
Level
0.40
1.5
Emissions, gm/mi
HC
3. 1
2.0
CO
30
7.1
NOX
0.66
1.5
Fuel
Economy
mpg
8.6
11.2
Remarks
w/o catalyst or
secondary air
w/o catalyst or
secondary air
Ford has conducted a considerable amount of single-cylinder
PROCO engine test work to determine the effects of various operating
parameters on emissions and fuel consumption (Ref. 8-10). The parameters
investigated include injection and ignition timing, overall air fuel ratio,
and EGR. As expected, retarded ignition and/or injection timing and EGR
reduce NO at the expense of higher HC, CO, and fuel consumption.
8.2.7.3 Fuel Economy
8.2.7.3.1 Texaco TCCS Engine
As discussed in Section 8.2.2. 1, an emission controlled M-151
TCCS vehicle was tested by the EPA in accordance with the 1975 FTP and after
8-39
-------
the vehicle had been subjected to a 50, 000-mile durability run (Ref 8-15).
The emission control system utilized on this vehicle consisted of EGR,
three catalysts, and intake throttling. It also required considerable main-
tenance work during the test period. The fuel economy, determined by
the EPA by means of the carbon balance method from six consecutive
tests, varied between 14.4 mpg and 17. 1 mpg with the average at 15.5 mpg.
Somewhat better fuel economy data have been reported by
Texaco for the standard emission control system and for modified configu-
rations (Ref. 8-8). These data, listed in Table 8-5, show that substantial
fuel economy and power losses are associated with the achievement of
extremely low NO emission levels. For example, a fuel economy loss
X.
of 33 percent and a power loss of 28 percent are incurred in achieving the
1976 Federal emission standards. Catalytic treatment of the exhaust gases
combined with moderately retarded combustion results in a 5 to 10 percent
fuel economy degradation (Table 8-5, case 11). As illustrated in Table 8-6,
similar results were obtained by Texaco for a TCCS-powered Plymouth
Cricket automobile equipped with EGR and oxidation catalysts and tested
with a 2, 500-pound inertia weight. The best fuel economy obtained with this
vehicle without EGR was 25.3 mpg. This compares with 20. 6 mpg for the
average 1973 model year Cricket certification vehicle tested at 2,250-pound
inertia weight (Ref. 8-18).
Level road fuel economy data for the naturally aspirated L-141
TCCS engine, installed in an M-151 light-duty vehicle, are presented in
Figure 8-14 (Ref. 8-8). The fuel economy of the TCCS-powered vehicle is
considerably better than the fuel economy of the vehicle with the standard
engine, especially at low vehicle speeds. For example, at 30 miles per
hour, TCCS has a 30 percent better fuel economy, and at 55 mph, the
difference is still 20 percent. The variations shown in Figure 8-13 for
the different fuels are directly related to the fuel heating value per gallon
and confirm the previously noted trends of specific fuel consumption
(Figure 8-12).
8-40
-------
D)
Q.
o
o
111
UJ
D
U.
STANDARD ENGINE
30 40
VEHICLE SPEED - mph
Figure 8-14. Level-Road Fuel Economy, M-151 Light-Duty Vehicle,
Texaco Naturally Aspirated, L-141 TCCS Engine
(Ref. 8-8)
8-41
-------
Except for vehicle speeds below about 30 mph, the level-road
fuel economy of the turbocharged M-151 TCCS vehicle is about 5 to 25 per-
cent better than that of the naturally aspirated TCCS engine, with the higher
improvement realized at vehicle speeds of 50 mph and above (Ref. 8-8).
Fuel economy tests over a closed-road course conducted by
TACOM on TCCS and standard-engine-powered M-151 vehicles support the
Texaco results. Based on the TACOM tests, the fuel economy of the TCCS
vehicle is between 39 percent and 73 percent better, depending upon the
average speed over the course (Ref. 8-8). Again this illustrates the
excellent part-load fuel economy characteristics of the TCCS engine.
8.2.7.3.2 Ford PROCO and Fast Burn Engines
The improved fuel-economy characteristics of open-chamber,
stratified charge engines with respect to conventional spark ignition engines
were demonstrated by Ford early in its development program (Ref. 8-9).
At that time, 430-CID FCP engines were installed in three vehicles and
tested at constant speeds between 30 mph and 70 mph. The data show a
28 percent to 34 percent improvement in the fuel economy of the FCP
vehicles, compared with similar standard vehicles.
Fuel economy data for the M-151 vehicle equipped with con-
ventional and PROCO engines and tested over the Federal Driving Cycle are
presented in Table 8-7. The values were computed from the measured
exhaust-gas composition utilizing the carbon-balance method (Ref. 8-10).
The fuel economy of the PROCO vehicle adjusted to meet the
1976 statutory emission standards at low mileage is about 24 percent better
than that of the conventional engine powered M-151 vehicle and about
12 percent better than the average 1973 model year vehicle tested at
2,750-pound inertia weight (Ref. 8-19). When adjusted for best efficiency,
the fuel economy of PROCO is approximately 36 percent better than the
average 1973 certification car tested at 2, 750-pound inertia weight. How-
ever, this improvement results in a substantial increase of the HC, CO,
and NO emissions.
8-42
-------
Fuel economy data obtained by Ford for its PROCO engine
equipped 1972 Montego vehicle are listed in Tables 8-8 and 8-9. The
computed fuel economy of this vehicle over the Federal Driving Cycle is
about 24 percent better than that of the average 1972 certification vehicle
in the 4, 500 pound weight class.
The fuel economy of the Fast Burn Phase I, 351-CID engine --
installed in a 1972 Torino automobile and tested according to the 1975 Federal
Test Procedure -- is listed in Table 8-10. As indicated, the vehicle meets
1976 emission standards at low mileage with a definite improvement in fuel
economy over a standard Torino. After 25,000 miles, the fuel economy of
the vehicle is essentially unchanged, although the HC and CO emissions have
increased considerably (Ref. 8-11).
As shown in Table 8-11, the fuel economy of Ford's Fast
Burn Phase II vehicle adjusted to low NO levels is not very impressive and
j£.
is equivalent to conventional 1973 Torino cars. However, the fuel economy
improves substantially by adjusting the NO emissions to the 1.5-g/mi level
X.
(Ref. 8-11).
8.2. 8 Noise
Engine noise, vibration, and harshness have been a problem in
Ford's early stratified charge engine installations. Subsequently, vehicle and
engine design modifications were incorporated, including insulation of the
vehicle floor pan and stiffening of a number of structural members on both
the engine and the vehicle. Although these changes resulted in a noticeable
improvement, the noise and vibration characteristics of these vehicles
remained unsatisfactory relative to standard engine/vehicle configurations
(Ref. 8-9).
No information is available on the noise and vibration levels
of the current Ford and Texaco SCE-powered vehicles.
8-43
-------
8.2.9 Smoke and Odor
Smoke measurements conducted by Ford on single-cylinder
stratified charge engines indicate exhaust opacities between 0.5 percent and
3 percent, compared with about 2 percent to 50 percent for diesels. In these
tests, the smoke level had a tendency to increase with increasing EGR
flow rate and engine load. A reduction in the smoke intensity could be
accomplished by increasing the squish action in the combustion chamber
to promote fuel mixing and vaporization (Ref. 8-9).
According to Ford (Ref. 8-9), the exhaust of its FCP engines
has an upleasant diesel-like odor which, in high concentrations, can cause
eye irritation. In general, the odor level is affected by the injection and
spark timing, decreasing with engine throttling and increasing as the EGR
flow rate increases (Ref. 8-9).
8.2.10 Driveability
During tests conducted by the EPA the emission-controlled
M-151 vehicle incorporating Texaco1 s TCCS engine has shown poor drive-
ability characteristics and insufficient power to negotiate the Federal Driving
Cycle. These deficiencies are attributed entirely to the emission control
system which was added to the engine to meet the 1976 statutory Federal
emission standards (Ref. 8-15). Conversely, the uncontrolled M-151
TCCS vehicle has exhibited very good driveability (Ref. 8-8). According
to Texaco, the cylinder injection results in rapid throttle response at all
operating conditions and eliminates the warmup period generally required
in carbureted engines. The uncontrolled engine also exhibits good starting
and acceleration characteristics, even at temperatures below 10 F, without
hesitation or stumble.
According to Ford, the driveability of its PROCO vehicle is
good and engine stumble has never been encountered, even at very low engine
speeds and full EGR flow. However, the acceleration performance of the
emission controlled vehicle is impaired, and the idle quality is not quite
8-44
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satisfactory with the short-electrode spark plugs. The driveability of the
vehicles utilizing Fast Burn engines was judged to be acceptable by Ford,
with the exception of the Phase II engine at high EGR flow rates (Ref. 8-11).
8.2.11 Current Status
For the past several years, Texaco and Ford, under contract
to the USA TACOM, have been involved in the development of stratified charge
engines for potential application in the military L-141 light-duty engine.
Without incorporation of emission-control equipment such as EGR, oxidation
catalysts, and intake throttling, the HC and NO emissions from the L-141
Jv
stratified charge engines are comparable to conventional spark ignition
engines, while CO is substantially lower. In this case, the fuel economy of
the stratified charge engine powered vehicles is of the order of 30 percent
higher than for equivalent conventional engines.
With emission control, the stratified charge engine powered
vehicles tested to date over the Federal Driving Cycle meet the 1976 statutory
emission standards at low mileage and show equal or slightly better fuel
economy than the average 1973 certification vehicles tested at the same
inertia weights. However, rapid deterioration of the emission control
system performance is a major problem area.
The stratified charge engines manufactured to date by Texaco
and Ford are hand-built experimental models, which are not suitable for mass
production. Specifically, the current cylinder-head design is not compatible
with mass-production techniques, and new production equipment and assembly
procedures are required for the manufacture of the fuel injection and air
fuel control systems.
A design study for a family of four, six, and eight-cylinder
PROCO engines has been recently completed by Ford, with parts commonality
being a prime consideration. Based on this study, Ford has concluded
that the PROCO engine concept can be incorporated in these engine
designs (Ref. 8-20).
8-45
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8.2.12 Projected Status
Ford feels that limited production of PROCO engines could be
feasible by 1977, providing the success experienced to date in the develop-
ment of the PROCO engine continues (Ref. 8-11). However, before pro-
duction commences, a number of potential problem areas must be further
evaluated, including engine-component durability and emission control system
deterioration characteristics, as well as fuel injection system requirements.
In addition, several of the engine components require design modification to
permit their manufacture by mass production techniques.
With respect to the Fast Burn concept, Ford plans to continue
the development of the Phase II configuration. In particular^ a new
improved induction system is needed to reduce the HC and CO emissions. It
appears that this concept might be sufficiently developed by 1977 to permit
low-volume production.
Since oxidation catalysts are required on Ford's and Texaco1 s
stratified charge engines to meet the 1975-76 statutory Federal emission
standards, the need for unleaded gasoline would be unchanged by the adoption
of these engine concepts. However, in view of the multifuel capability of
the Texaco engine and its insensitivity to octane and cetane number, utiliza-
tion of this engine would alleviate potential problems related to removal
of lead from the fuel.
8.3 DIVIDED-CHAMBER STRATIFIED CHARGE ENGINES
8.3.1 General
Historically, the concept of a divided combustion chamber
can be traced back to the first "oil engines" in the pre-diesel era. These
engines were operated on heavy oils or naphtha injected into a spherical
prechamber and vaporized in contact with the red-hot walls of the thermally
insulated prechamber. Later, attempts were made to apply this concept
to the spark ignition gasoline engines, as documented by over 100 patents.
However, a practical and widespread application of the prechamber concept
8-46
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was realized only in the high-speed diesel engines, primarily due to the
pioneering efforts of Ricardo in the early 1920s. The announcement in
1972 of the development of Honda's Compound Vortex Controlled Combustion
(CVCC) engine marks the first practical prechamber spark ignition engine
designed for use in automotive mass-production applications.
The combustion process in a divided-chamber engine proceeds
in two stages. First, a rich air fuel mixture is ignited in the prechamber.
Subsequently, the burning charge expands from the prechamber to the main
combustion chamber, through communicating openings and ignites the lean
air fuel mixture or the pure-air charge inducted into the main chamber.
Several controlling factors are involved in this two-stage combustion process.
These include the type of fuel and air-supply system utilized; the relative
size of the prechamber to the main chamber; and the geometry of the
prechamber, main chamber, and communicating passages.
Compared -with open-chamber engines, the divided-chamber
configurations have certain inherent disadvantages that are related to the
higher surf ace-to-volume ratio of the combustion chamber and the higher
flow-turbulence level during combustion. As a result, the heat losses to
the cooling jacket and the emission of unburned hydrocarbons are increased.
The net improvement in engine emissions and/or fuel economy depends on
how well these effects are balanced by better combustion in the main chamber
and improvement in engine cycle efficiency (for unthrottled engine operation).
Many domestic and foreign investigators and organizations are,
or have been working on the concept of the divided-combustion-chamber gaso-
line engine. In addition to the most advanced development activity conducted
by the Honda Motor Company, research and development work have been
performed by the Ford Motor Company, Phillips Petroleum Company,
Stanford University, University of Rochester, University of Wisconsin,
Dynatech Corporation, and Volkswagen.
8-47
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8.3.2 Power Plant Description
The divided-combustion-chamber designs currently known can
be divided into three classes according to the ratio of the prechamber volume
to the total compression-cylinder-clearance volume (prechamber, connect-
ing passage, and main chamber): miniature prechambers, medium-size
prechambers, and large prechambers. The design features of these engine
categories are briefly described in the following paragraphs; discussions
of selected engine designs are presented in Sections 8.3.2. 1 through 8. 3. 2. 3.
The miniature prechamber engines employ a small pre-
chamber generally formed by an enclosed cavity around the spark plug elec-
trodes of each cylinder. In these designs the prechamber volume is less
than 10 percent of the total compression volume. The principal function of
the miniature prechamber is to intensify ihe ignition and to stabilize the
combustion process of the slightly lean air fuel mixture in each engine
cylinder. This is achieved primarily by increased turbulence rather than
stratification of the air fuel mixture. The effectiveness of this concept for
emission reduction is rather moderate and might be attractive only in con-
junction with other emission control methods, such as exhaust gas recircu-
lation. This engine concept is exemplified by the Ford Torch engine and by
the designs patented by Jozlin and Bishop.
The medium-size prechambers are the favored development
targets of the various prechamber combustion engine concepts conceived
to date (over 50 patents since 1910). These prechambers rarely exceed
25 percent of the total combustion chamber volume. The principal function
of the medium-size prechamber is to (1) provide a means for effective
stratification of the air fuel mixture, (2) facilitate ignition of the lean mix-
ture in the main chamber, and (3) help generate the flow pattern and
turbulence conducive to optimal combustion in the prechamber and in the
main chamber. Engines in this category include the Honda CVCC engine
and the designs by Broderson-Conta, Dynatech, Heintz, Nilov, Phillips,
and Volkswagen.
8-48
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The large-prechamber engine configuration represents a
more recent concept which has been researched by Newhall and by the Ford
Motor Company. In this design, the size of the prechamber ranges from
about 50 to 85 percent of the clearance volume and communicates with the
main chamber (secondary chamber) through a large orifice. The objective
of the large prechamber is to provide for reliable stratification of the air
fuel mixture. Similar to divided-chamber dies els, all fuel is injected into
the prechamber. Upon ignition, the burning charge expands from the pre-
chamber into the main chamber to continue the combustion process. It is
hypothesized that the relatively cool air in the main chamber assists in
quenching the NO formation reactions while providing additional oxygen
required for the completion of oxidation reactions of carbon monoxide and
hydrocarbons.
The aforementioned functions are not distinctly limited to
only one size category. A certain overlap will result, depending upon the
size and geometry of the prechamber and the prechamber fuel and air supply
systems utilized.
8.3.2.1 Miniature Prechamber Engines
8.3.2.1.1 Ford Torch Ignition Engine
The Ford Torch Ignition Engine illustrated in Figure 8-15
features a torch chamber cavity or miniature prechamber located outside
the main combustion chamber with which it is connected by one or more
small holes. The carbureted air fuel mixture is introduced by a conventional
manner into the engine cylinder, and part of it is transferred into the torch
chamber during the compression stroke. The spark plug, located in the
torch chamber, ignites the mixture in this chamber and "torches" of burning
mixture are expelled through the torch chamber holes into the main com-
bustion chamber. The torches of burning mixture penetrate deep into the
main chamber, thus creating large area flame fronts that promote rapid
completion of the lean-mixture combustion process.
8-49
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TORCH CHAMBER
TORCH CHAMBER ORIFICES
COMBUSTION CHAMBER
Figure 8-15. Ford Torch Ignition Engine (Ref. 8-11)
8-50
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Since there is no provision for enrichment of the combustible
mixture in the torch chamber, the air fuel ratio of the engine is adjusted to
about 15.5 (Ref. 8-11). The design changes with respect to conventional
carbureted gasoline engines are limited to the engine head in the vicinity of
the spark plug. No change in weight, volume, materials, or fuel requirement
is indicated (Ref 8-11).
8.3.2.1.2 Other Concepts
A number of recent patents involving the concept of the
miniature prechamber indicate current interest, research, and development
in this engine configuration. Two of these concepts, which present new ideas
for further functional improvement, are discussed in the following paragraphs.
Figure 8-16 presents the concept of a miniature prechamber
equipped with a check valve according to a U. S. patent granted to Joseph A.
Jozlin (Ref. 8-21). The objective of the check valve arrangement is to pro-
vide a sufficiently large port for the unobstructed entry of combustible mix-
ture into the prechamber during the compression stroke and, subsequently,
to block free exit of the burning mixture from the prechamber at the beginning
of the piston power stroke. Upon ignition, high pressure is produced in the
prechamber, and the burning mixture is then forced at high velocity through
a set of connecting holes into the main combustion chamber. In an
alternative embodiment, the check valve is replaced by a cam-operated
poppet valve.
Figure 8-17 shows the miniature prechamber concept patented
by I. N. Bishop, et al. (Ref. 8-22). In this case, direct fuel injection is
employed instead of carburetion. The fuel injection nozzle is disposed to
inject fuel into the main combustion chamber contiguous to the opening
between the prechamber and the main chamber. The configuration of the
main chamber and the positioning of the connecting hole (in the prechamber)
are such that an air flow is set up from the main combustion chamber into
8-51
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SPARK PLUG
£\ X A X A Sj
•S/--rr=zr- -/J >
LA, ^—— - =r=-f/y
PRECHAMBER
Figure 8-16. Jozlin Prechamber Concept(Ref. 8-21)
SPARK PLUG
Figure 8-17. Bishop Prechamber Concept (Ref. 8-22)
8-52
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the prechamber during the compression stroke of the piston. The air flow
sweeps a portion of the fuel droplets injected into the prechamber to provide
a locally rich air fuel mixture. Since the mixture formed in the prechamber
is richer than exists in the main chamber, it would appear that this concept
has the potential of further extending the lean limit of the engine operation.
8.3.2.2 Medium Size Prechambers
8.3.2.2.1 Honda CVCC Engine
The original Honda CVCC (Compound Vortex Controlled
Combustion) engine is an in-line, four-cylinder, four-stroke, 1948 cm dis-
placement (119 in ) engine with a rating of 70 bhp at 5,000 rpm (Ref. 8-23).
Prechamber volume is in the range of 1 0 to 20 percent of the total combustion
chamber volume, and both parts of the chamber communicate through a small
orifice. Figure 8-18 shows the principle of the design.
In this design, the smallest venturi of the three-barrel carburetor
supplies a rich mixture to each prechamber. The other two Venturis supply the
main engine chambers with a very lean mixture. The main chamber utilizes a
conventional inlet valve located in the chamber, while the mixture to the pre-
chamber enters through a small inlet valve that opens into the prechamber.
The prechamber inlet valve is actuated by means of an independent small
rocker arm driven by the engine camshaft. The charge is stratified into
three air fuel ratio regions: (1) a rich region in the prechamber, (2) a lean
region in the main chamber, and (3) a region of moderate mixture strength
in the vicinity of the orifice connecting both chambers. Combustion is initiated
by the spark plug located in the prechamber.
According to Honda, combustion in the CVCC engine is controlled
and proceeds at very slow rates to reduce the peak combustion temperature
and the formation of NO. However, the temperatures are maintained high
enough for a sufficiently long period of time during the expansion stroke to
complete the oxidation of CO and HC. Conventional unmodified spark plugs
are used on this engine. Apparently, optimum values exist for this engine
with respect to configuration and size of the prechamber and the relative
8-53
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Figure 8-18. Honda CVCC Divided Chamber Stratified
Charge Engine (Ref. 8-23)
8-54
-------
positions of the auxiliary intake valve, spark plug, torch nozzle, and main
combustion chamber. The air fuel ratios supplied to the prechamber and
the main chamber vary with the operating conditions and are controlled with
respect to prechamber scavenging, desired emission levels, and other
performance parameters.
8.3.2.2.2 Broderson-Conta Engine (University of Rochester)
Figure 8-19 presents the concept of the Broderson-Conta
engine (Ref. 8-24). Supposedly, this engine approches the diesel efficiency
at light loads. In the engine, the fuel is directly injected into the prechamber
while an unthrottled, carbureted, lean air fuel mixture is inducted into the
main chamber. The engine load is varied by changing the amount of fuel
supplied through the carburetor, and at very light loads the main chamber is
charged with air only. Apparently, engine noise has been a persistent
problem area.
8.3.2.2.3 Dynatech Prechamber Engine
A rather sophisticated approach was followed by Dynatech
Corporation, which developed a prechamber concept for small, two-stroke
engines (Ref. 8-25). In this design, fuel is injected at low pressures directly
into both the prechamber and main chamber. The rate of fuel injection into
the prechamber remains constant, while the main chamber-injection rate is
varied with engine load. The injection system is electronically controlled,
and air admission is unthrottled.
8.3.2.2.4 Ford Three-Valve-Prechamber Engine
Ford has conducted Three-Valve-Prechamber engine research
and development work using both single-cylinder and V-8 engines (Ref. 8-11).
In general, development-vehicle emission results have not reached the levels
predicted from dynamometer tests. Major shortcomings of the present hardware
include insufficient heating of the main and secondary induction systems, which
results in high HC emissions due to excessive richness produced during vehicle
deceleration. Vehicles equipped with current hardware can be operated satis-
factorily with a fixed throttle angle in the secondary rich-mixture carburetor.
No design information has been released by Ford.
8-55
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PRECOMBUSTION
CHAMBER
COMBUSTION
CHAMBER
FUEL INJECTOR
THIRD
VALVE
SPARK PLUG
GAS PASSAGE
Figure 8-19. Broderson-Conta Engine (Ref. 8-24)
8.3.2.2.5 Heintz Ram-Straticharge Engine (Stanford University)
As illustrated in Figure 8-20, the Heintz engine is similar to
previously discussed concepts although utilizing a more complex prechamber
configuration (Ref. 8-26). Like the Honda CVCC design, the Heintz engine
employs intake air throttling to limit the maximum air fuel ratio to values
below 30 to assure complete combustion of the fuel. The original Heintz
engine has since been modified and applied to both two and four-cycle
engines. No data are available on the performance and emissions of these
designs.
8.3.2.2.6 Nilov Engine
The basic concept of the Nilov engine is similar to the Honda
CVCC system (Ref. 8-27). However, as shown in Figure 8-21, there appears
to be a significant difference in the geometry of the main chamber and the
location of the inlet valve. Also, the volume of the prechamber appears
substantially smaller. Apparently, the Nilov engine has been developed to a
high state of perfection and seems to be ready for production (Ref. 8-27).
8-56
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FUEL LINE
THIRD VALVE
PRECOMBUSTION
CHAMBER CC\I
GAS PASSAGE
PLUG LEAD
SPARK PLUG
Figure 8-20. Heintz Ram Straticharge Engine
(Ref. 8-26)
LEAN MIXTURE
CARBURETOR
/oAx
-------
8.3.2.2.7 Phillips Petroleum Prechamber Engine
A comprehensive basic study was conducted by the Phillips
Petroleum Company on a CFR engine equipped with a retrofit prechamber
(Ref. 8-28). In this design, the prechamber is screwed into the standard
spark plug hole, and the spark plug is then attached to the prechamber
which is equipped with an auxiliary inlet valve.
The volume of the prechamber is approximately 13 percent
of the total compression-clearance volume. In tests performed by Phillips
at speeds between 1,000 and 1,800 rpm, the power was controlled by
throttling the premixed air fuel charge into the engine cylinder. A very
rich fuel air mixture, containing five times the stoichiometric fuel, is
inducted into the prechamber at a rate of 2 percent of the main-chamber
mixture flow.
8.3.2.2.8 Volkswagen Prechamber Engine
Volkswagen conducted considerable research and development
on prechamber engines. Its latest development is illustrated in Figure 8-22.
In this design, a spherical prechamber utilizing approximately 25 percent
of the compression volume, is connected to the main combustion chamber
by a flow-transfer passage. The main combustion chamber is disc shaped
and contains no squish surfaces. The prechamber has no auxiliary valve.
Fuel is injected into the prechamber by means of a low-pressure injection
nozzle. The injection nozzle and spark plug are arranged in such a manner
that a relatively rich mixture is present at the spark plug under all operating
conditions. Mixture preparation is assisted by the shape of the prechamber,
with the transfer passage entering almost tangentially. Load control is
accomplished by the amount of carbureted mixture introduced into the main
combustion chamber.
According to Volkswagen, fuel injection timing need not be
varied over the entire load and speed range. This represents a considerable
design simplification, compared with the mixture stratification process in
J-51
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Figure 8-22. Volkswagen Prechamber Engine (Ref. 8-29)
an open combustion chamber, which generally requires the use of an injec-
tion timing control system. Initially, tests were conducted exclusively
without intake air throttling. Later, partial throttling was employed in order
to reduce the emissions at low engine loads.
8.3.2.3 Large Prechambers
8.3.2.3.1 Newhall Engine
The divided combustion chamber concept by Newhall and
El-Messiri is shown in Figure 8-23. In this design, all fuel is injected into
the primary chamber at an intermediate point during the compression stroke.
The geometry of the combustion chamber and the injection timing are so
arranged that no raw fuel enters the secondary chamber. The spark plug,
which ignites toward the end of the compression stroke, is located in the
primary chamber near the orifice connecting both chambers.
8-59
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PRIMARY
CHAMBER
N
/'{f
xji
#
'/.'
"-S
\
-<>
1
SECONDARY
CHAMBER
TO FUEL
METERING
SYSTEM
_f
Figure 8-23. Newhall Prechamber Engine
(Ref. 8-30)
The power output of the engine is varied by conventional throttling.
Air and fuel throttles are controlled simultaneously to maintain an air fuel
mixture of appropriate strength in the primary chamber under all load conditions.
Briefly, combustion in the Newhall engine proceeds as follows. Upon ignition
of the prechamber mixture, combustion gases expand immediately into the
secondary chamber, which initially contains relatively cool air. In principle,
the rapid mixing and cooling process that ensues minimizes the formation of
oxides of nitrogen, while at the same time the excess air contained in the
secondary chamber promotes complete oxidation of hydrocarbons and carbon
monoxide.
8-60
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8.3.2.3.2 Ford Large Prechamber Engine
In addition to the work of Newhall, the Ford Motor Company
has reported research and development work on large-volume prechamber
engines (Ref. 8-11). Tests on a single-cylinder engine were conducted with
and without EGR. With optimized adjustment under steady-state conditions,
encouraging emission data were obtained. However, the operating range
of the engine was limited to light-to-medium loads due to combustion
harshness and poor air utilization.
In order to facilitate further development of this concept on
multicylinder engines, a design effort has been initiated by Ford to incorporate
this combustion concept into the 400 CID V8 engine. Dynamometer testing of
this engine -was scheduled to start in November 1973.
8.3.3 Performance
8.3.3.1 Emissions
8.3.3.1.1 Honda CVCC Engine
Evaluation of the emission characteristics of the Honda CVCC
engine was conducted by the EPA (Refs. 8-31 and 8-32) and by the Honda Motor
Company (Ref. 8-23). In the EPA test program, three Honda Civic cars
equipped with CVCC engines were tested according to the 1975 Federal Test
Procedure, using an inertia weight of 2,000 pounds. Two of the vehicles
had odometer readings of about 1, 500 miles, and the third car had just com-
pleted a 50,000-mile durability run. A summary of these emission tests is
presented in Table 8-12, showing that the emissions from the vehicles
were consistently below the 1975 Federal emission standards.
One of the three cars was also tested with an inertia weight
of 3,000 pounds. The vehicle had adequate power to keep up with all the
accelerations of the Federal Driving Cycle. Because of time constraints,
this was a "hot start" test; i. e. , the start of the test did not follow the
standard 12-hour vehicle soak period but began shortly after the car had been
8-61
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Table 8-12. Emissions and Fuel Economy, Honda
CVCC Vehicles (Ref. 8-33)
Vehicle
Low mileage car
(#3652. Five tests.)
Average
Maximum3-
Minimuma
Low mileage car
(#3606. One test.)
50, 000-mile car
(#2034. Four tests.)
Average
Maximum
Minimum
Emissions, gm/mi,
1975 FTP
HC
0.18
0.21
0.15
0.23
0.24
0.26
0. 19
CO
2.12
2.28
1.96
2.00
1.75
1.85
1.70
NOX
0.89
1.05
0.75
1.03
0.65
0.73
0.57
Fuel Economy, mi/ gal
1975 FTP
22. 1
22.4
21.9
20.7
21.3
22.2
20.8
1972 FTP
21.0
21.5
20.6
19.5
19.8
20.0
19.5
FTP = Federal Test Procedure (grams per mile)
The values of the different columns do not necessarily correspond to the
same test.
tested at 2,000-pound inertia weight. Although the initial period of the
"cold start" was omitted, the test results appeared to show that the emis-
sions of a Honda vehicle of 3,000 pounds, equipped with the same CVCC
engine, would be lower than the statutory 1975 emission standards
(Ref. 8-31).
Particulate emission tests conducted by Dow Chemical
Company under EPA contract indicate that the particulate emissions from
the CVCC vehicles are comparable to those from convenient engines using
equivalent fuels .
8-62
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To date, Honda has conducted emission durability tests on
seven 2.0-liter CVCC vehicles accumulating a total of 437, 000 miles. Based
on these tests, the deterioration factors are approximately 1. 10 to 1. 15 for
HC, 1.05 to 1.10 for CO, and 1. 00 to 1. 05 for NO , showing good emission
^c
control durability of the CVCC engines.
Honda has adapted the CVCC process to larger engines to demon-
strate the feasibility of the concept in full-size American cars. Table 8-13
presents Honda's emission test results for a Chevrolet Vega car equipped with
a 140 CID CVCC engine. As indicated, the emissions are below the statutory
1975 standards, and the fuel economy is slightly better than for the non-
modified Vega.
Emission data obtained by the EPA for a CVCC modified 350 CID
Chevrolet Impala automobile are listed in Table 8-14 (Ref. 8-32). Tests 1 and
4 demonstrate that CO and HC emissions below the statutory 1975-76 levels
can be achieved with this vehicle, and that oxides of nitrogen levels are con-
sistently below the 1976 interim standard of 2.0 grams per mile. During
tests 2 and 3, problems associated with the hot soak period resulted in high
levels of CO and HC during the bag 3 portion of the tests. The high CO level
in bag 3 of test 2 appeared to be associated with flooding of the prechamber
carburetor. This was corrected by Honda technicians prior to test 3. The
high hydrocarbon level in bag 3 of test 3 was apparently caused by a false
hot start which resulted in excessive cranking by the EPA driver. Another
false start occurred on test 4, but the more experienced Honda driver was
able to recover with less hydrocarbon penalty. It is believed that the hot
start problems experienced on the prototype vehicle can be corrected with
further engineering effort.
The consistently low cold start (bag 1) hydrocarbon and CO levels
during all of the tests should be emphasized. An average value for bag 1 of
0.38 gm/mi HC and 3.51 gm/mi CO was observed. Good control of oxides of
nitrogen was also demonstrated with this vehicle. An average value equal to
1.72 gm/mi was obtained for the 1975 FTP testing.
8-63
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Table 8-13. Low-Mileage Emissions and Fuel Economy of
Vega Vehicles Modified for CVCC (Ref. 8-23)
CVCC - 140 CID
CVCC - 140 CID,
improved
CVCC - 140 CID,
improved
Original 140 CID
Test
Date
9-72
11-72
4-73
-
Emissions, £m/mi
HC(1)
0.26
0.26
0. 22
2. 1
(1)1975 FTP
(2)1972 FTP
co'1'
2.9
2.6
2.4
10.6
NO*1'
X
1.2
1.2
1.2
3.8
Fuel
Economy
mi/gal(2)
17-9
18.6
18.9
17.2
No.
of
Tests
2
2
3
2
Table 8-14. Low-Mileage 1975 FTP Emissions and Fuel Economy of
Impala Vehicles Modified for CVCC (Ref. 8-32)
Test No.
1
2
3
4
HC
(gm/mi)
0.27
0.23
0.80
0.32
CO
(gm/mi)
2.88
5.01
2.64
2.79
NOX
(gm/mi)
1.72
1.95
1.51
1.68
Fuel Consumption
(mi/gal)
10.5
11.2
10.8
10.2
Note: Above tests run at 5,000-pound inertia and 14.7 rear wheel
hp at 50 mph (hp includes 10 percent for air conditioning)
8-64
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8.3.3.1.2 Other Concepts
The emission data available for the other prechamber
engines are briefly discussed in the following paragraphs.
Hot-start emission and fuel economy data published by Ford
for its Torch Ignition Engine are presented in Table 8-15 (Ref. 8-11). The
test vehicle was a 4, 500-pound 1972 Gran Torino Sport equipped with an
8.5 compression ratio, 351-CID engine, and modified for torch ignition.
Unfortunately, no baseline emission data from an unmodified engine are
shown to indicate the degree of emission reduction attained. The NO levels
x
are interesting; however, the relatively low CO values may be due primarily
to the air injection into the exhaust manifold used in conjunction with the
torch ignition concept. All reported tests were run with EGR, and the NO
2C
and CO emissions are typical of the levels obtained from conventional
engines equipped with EGR and with lean carburetor settings.
Hot-start emission data obtained by Ford for a 1973 LTD and
a 1973 Gran Torino equipped with 400-CID Three-Valve-Prechamber Engines
are presented in Table 8-16. As indicated, CO is very low and comparable
to the Honda CVCC data, but HC is much higher. Because of induction sys-
tem heating problems, only hot-start tests were performed by Ford on these
vehicles. Preliminary emission data taken by Ford on a single-cylinder
experimental engine incorporating a large volume prechamber are listed in
Table 8-17.
The Dynatech concept was tested by the EPA in a Saab vehicle
in accordance with the 1975 FTP (Ref. 8-25). In these tests, the emissions
were 6.4 g/mi HC, 3.6 g/mi CO, and 0.3 g/mi NOx> The HC emissions
are quite high and exceed the 1976 standards by an order of magnitude. Con-
versely, the NO level appears to be safely below the standard, while CO
5t
approaches the standard. Since JP-4 was used in these tests, there is some
question as to the effect of this fuel type on the emission results.
The effect of the Phillips Petroleum retrofit prechamber on
the emissions of an experimental CFR engine is presented in Figure 8-24.
8-65
-------
Table 8-15. Hot-Start Emissions and Fuel Economy, Ford Gran
Torino with Torch Ignition Engine (Ref. 8-11)
Test
No.
5
8
10
12
13
Emissions, gm/mi
HC
2.78
3. 13
2.72
1.86
2.00
CO
8.75
5.67
5.73
0.64
13.87
NO
X
2.22
1.57
2.26
1.27
1.26
Fuel^1)
Economy
(mi/ gal)
12.25
11.55
11.73
11.76
11.71
%
EGR
7-15
8-15
8-16
5 - 15
6 - 15
Remarks
15.5:1 A/F TDC timing
15.5:1 A/F w/air
injection TDC timing
15.5:1 A/F 6° BTDC
timing w/air injection
15. 5:1 A/F TDC timing
w/air injection; 1 hole
torch
15.5:1 A/F TDC timing
w/air injection; 1 hole
torch
Carbon balance method; Federal Driving Cycle
Table 8-16. Ford Vehicle Hot-Start Emissions and Fuel Economy,
Three-Valve-Prechamber Engines, 1975 FTP(Hot)
(Ref. 8-11)
Test
Vehicle
1973 LTD*
1973 Gran ;,:
Torino
Engine
400 CID
Phase 1
400 CID
Phase 2
Emissions, gm/mi
HC
1.78
2.40
CO
3.36
3.92
NO
X
1.29
1.25
Fuel
Economy
(mi /gal)
9.8
11.0
EGR
Flow
0
0
"inertia weight 4, 500 Ib
8-66
-------
Table 8-17. Ford Large-Prechamber Emissions and Fuel
Economy, Single-Cylinder Engine (Ref. 8-11)
Speed
(rpm)
1500
1500
Indicated
Mean Effect.
Pressure, psi
40
70
i_ Emissions, ^r/ihp-hr
HC
1.2
1. 1
CO
5.0
6.0
NO
X
0.5
0.5
Indicated Fuel
Consumption
(Ib/ihp-hr)
0.395
0.377
On.50
U-.C
V)'-
~>40
56
^48-
Q. 40
|32
o>
o-24
°16
i.-4-xt--r
40 60 80 100120140
IMEP, psi
s_
"i.20
10
o
.C
Q.
24
20
1 16
°! 12
x
§ 8
—»—t
40 60 80 100 120140
IMEP, psi
Note: Standard Engine (CR = 6.0)
_ - - Prechamber Engine (CR = 6.46)
1000 rpm
Maximum Power Spark Timing
Figure 8-24. Phillips Single-Cylinder Prechamber Engine Emissions and
Indicated Specific Fuel Consumption vs Indicated Mean
Effective Pressure (IMEP) (Ref. 8-28)
8-67
-------
The prechamber-equipped engine produces less NO over the entire range
of engine load (IMEP). NO emission was very low for engine loads below
Ji
70-psi IMEP. The comparison with emissions from an equivalent standard
engine demonstrates the potential of the prechamber engine for reducing
NO emissions. However, emissions of HC are higher in this prechamber
engine at loads greater than 70-psi IMEP and markedly lower for engine
loads below 60-psi IMEP. A similar trend was observed for CO emission.
Reference 8-27 presents emission test results for a wide range of engine
loads, rpm, air fuel equivalence ratios, and percent of throttle opening.
Steady-state emissions published by Volkswagen for its
single-cylinder prechamber test engine are presented in Figure 8-25 as a
function of mean effective pressure. As indicated, the size and shape of
the connecting orifice have some effect on the emissions. The HC emissions
are quite high, but Volkswagen feels that significant improvements could be
made by means of design modifications (Ref. 8-26).
Emission data from Newhall's single-cylinder, large-
prechamber engine are plotted in Figure 8-26. The emission levels are
quite low, in particular the HC at equivalence ratios of about 0. 8. However,
the results are limited to only one set of engine operating conditions.
Furthermore, from the strong effect of fuel air equivalence ratio on the
emission of all three contaminants, it could be expected that the emissions
throughout the entire operating range of the engine will be very sensitive
to any deviation of the operating parameters from the optimal adjustment.
8.3.3.2 Fuel Economy
The available fuel economy data for prechamber engines are
briefly discussed in the following paragraphs.
Average, maximum,and minimum fuel economy data reported
by Honda for its CVCC-powered test vehicles are listed in Table 8-12. As
shown, the average 1975 FTP fuel economy of the three Honda vehicles tested
(21.3 mpg) was 16 percent lower than the average of the 2,000-pound 1973
certification vehicles tested by EPA (25.5 mpg). Honda data comparing
8-68
-------
oo
i
SO
1000
800
a.
a.
0"
o
600
400
200
SHAPE OF ORIFICE:
a ROUND 78.5 rW
A OVAL 86.5 mm*
O OVAL 123.0 mm2
105'
2345
MEAN EFFECTIVE PRESSURE
6 bar
Figure 8-25. Emission Characteristics, Volkswagen Single-Cylinder
Prechamber Engine (Ref. 8-29)
-------
at
i
LJ
O
to
w>
2
UJ
0.01
0.008
0.006
0.004 —
00002
Note:
1600 rpm
Full Throttle
MBT Ignition
0.5 0.6 0.7 0.8 0.9
FUEL AIR EQUIVALENCE RATIO
1.0
Figure 8-26. Preliminary Emission Data, Newhall
Prechamber Engine (Ref. 8-33)
8-70
-------
the CVCC-powered Civic to the standard Civic with a conventional engine
also show about 10 percent fuel economy penalty for the CVCC version.
However, the fuel economy of this car at 3,000-pound inertia weight
(18. 7 mpg) compared quite favorably with both of the average 3,000-pound
1973 certification vehicles (16. 2 mpg). Comparison of the data in Tables 8-13
and 8-14 indicate slightly better fuel economy for the CVCC-modified Vega
and Impala vehicles.
Fuel economy data from Ford's 1972 Gran Torino vehicle
equipped with a Torch Ignition Engine and EGR are listed in Table 8-15. As
indicated, the average fuel economy of this vehicle operated with different
EGR rates and spark timing is about 10 percent better than that of the average
1972 certification vehicle tested at 4,500-pound inertia weight.
Fuel economy data for Ford's Three-Valve Prechamber
Engine installed in a 1973 LTD and a 1973 Gran Torino vehicle are listed in
Table 8-16. As indicated, the fuel economy of the two vehicles is slightly
better than for the corresponding 1973 certification vehicles.
Limited indicated specific fuel consumption (ISFC) data for
Ford's single-cylinder large prechamber engine were presented in Table 8-17.
The data are in reasonable agreement with the fuel consumption data reported
by Newhall for a similar type engine (Ref. 8-30).
The specific fuel consumption data plotted in Figure 8-24 for
the Phillips prechamber engine indicate better economy relative to the stan-
dard CFR engine for indicated mean effective pressures (IMEP) below about
70 psi. Conversely, above 70 psi, the standard engine is better. Con-
sidering the low power requirements in urban traffic, this result suggests
the possibility of substantial fuel savings that might be realized with the
prechamber engine.
8.3.4 Noise and Odor
Engine noise and roughness of operation tend to be more
severe in prechamber engines than in conventional spark ignition engines.
This is particularly noticeable at higher engine loads. Some success in noise
8-71
-------
suppression was achieved by modifications in the geometry of the
communicating passages between the prechamber and the main chamber
(Ref. 8-23). Currently, it is not clear whether the engine noise and rough-
ness are due to high pressure rise rates in the engine cylinder or some
secondary gas dynamic effects in the main chamber.
There appears to be disagreement regarding the exhaust odor
characteristics of prechamber engines. Honda (Ref. 8-23) reports aldehyde
levels below 0.03 grams per mile for CVCC-powered vehicles, much lower
than normally found in conventional spark ignition engines. Test data pub-
lished by Wimmer (Ref. 8-28) indicate that the aldehyde emissions in
prechamber engines increase with increasing overall air fuel ratio.
8.3.5 Maintenance Requirements
No special maintenance problems or needs have been identi-
fied with any of the prechamber engine concepts presently in development.
Honda, after extensive durability tests, concluded that the maintenance pro-
cedures required for CVCC engines are no different from those for conven-
tional spark ignition engines, and the emissions can be preserved within the
scope of ordinary maintenance procedures.
8.3.6 Driveability
The only information on vehicle driveability effects related to
the incorporation of a prechamber engine was reported by the EPA. Appar-
ently, no problems were encountered during road performance evaluation of
the CVCC vehicles; the engines were very responsive and showed good
acceleration capability (Ref. 8-31).
8.3.7 Current Status
The development status of the previously discussed spark
ignition prechamber engine concepts is summarized in Table 8-18. To date,
the Honda CVCC engine is the only prechamber engine which has been sub-
jected to extensive performance tests and evaluations and which has achieved
production prototype status. Apparently, a number of the other concepts in
8-72
-------
Table 8-18. Current Prechamber Engine Development Status
o
4-1
nj
'c
E
d
•3
V
S
rt
1 1
M —'
(U M O
CX D-<
•^ ^ ^
U W ° C
aj !n ^ 'zi
1 -S P^ ^
J .S c -3
>, 7^ o o
i O tj CX
>*-( 14H Cl ^>. !0 ^
O O 3 4J ^ cr
c c c 5 9 (S
0 0 o £> ^
1 S ° , - S j; a
.2 " -^ g ^ rt
G S rH ^ O ^-1 O U
WW^OEQCXO
Performance
N
N
N
N
2
N
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N
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o
4-J
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ti
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30%)
-5 radically decreased ( > 80%)
J-73
-------
the medium-size category also have reached an advanced state of
development. Conversely, little concrete information is available relative
to the status in research and development of small and large prechamber
engines. Based on the technical information available, the following obser-
vations can be made:
a. A substantial reduction of the NOX, HC, and CO emissions
is reported for the medium-size prechamber concepts
discussed in this report.
b. In several cases, the fuel consumption of the prechamber
engine is higher than that of equivalent conventional spark
ignition engines. However, this may be compensated for
to some extent by an attendant reduction in fuel-octane
requirement.
c. Relative to conventional spark ignition engines, prechamber
engines tend to have higher noise levels and may suffer
some loss in power.
d. The production cost of prechamber engines is expected to
be slightly higher than for conventional engines.
e. No driveability problems are indicated.
f. Good engine durability and no unusual maintenance problems
and expenses are expected for mass-produced prechamber
engines.
8.3.8 Projected Status
The trends noted above may change significantly if unthrottled
engine operation can be achieved in any of the discussed concepts. In addi-
tion to lowering emissions, this could result in an appreciable reduction in
fuel consumption.
Two of the reported concepts (Broderson-Conta and Dynatech
Corporation) are designed for unthrottled operation. However; the lack of
emission and performance data precludes the assessment of the feasibility
of these engines at this time.
Even if future research should show that the condition of
unthrottled operation cannot be satisfied, it might be possible to develop an
optimal prechamber concept which would achieve a substantial reduction
8-74
-------
in emissions without appreciably increasing fuel consumption. In exploring
this possibility, the following considerations may prove to be valid:
a. Miniature Prechambers. With decreasing size of the
prechamber, the engine characteristics approach the
characteristics of a standard engine. As a result, the
potential for emission reduction decreases, but the
fuel economy penalty decreases also.
b. Large Prechambers. With increasing size of the pre-
chamber, the engine characteristics approach the charac-
teristics of an open-chamber stratified charge engine.
The potential for emission reduction increases; however,
the typical problems associated with open-chamber
stratified charge engines become more pronounced.
c . Medium Size Prechambers
1. The benefits of emission reduction and fuel economy
improvement might be realized at low engine loads
only. At high engine loads, these trends might be
reversed.
2. Achievement of an optimal engine system may
require development of more sophisticated
prechamber concepts.
8.4 CONCLUDING REMARKS
To date, no information has been published on studies related
to two-stage combustion in prechamber gasoline engines. Hypotheses on the
kinetics of NO formation in the prechamber, transition zone, and main
chamber have been formulated (Ref. 8-30), and certain flow patterns in the
prechamber and main chamber have been assumed by most investigators
(Ref. 8-23). However, these hypotheses and assumptions have not been
substantiated by theoretical and/or experimental investigations. In reality,
the stratification pattern anticipated from a given chamber design may be
distorted by gas dynamic (e.g. , resonance) effects in both prechamber and
main combustion chamber to such a degree that the optimal combustion con-
ditions are not attained (Refs. 8-34 and 8-35). These effects may also be
associated with the noise and roughness of prechamber-engine operation.
To elucidate complex behavior of this kind, comprehensive basic studies
appear indispensable.
8-75
-------
The conclusive assessment of the potential of the gasoline
engine prechamber concept for offering a satisfactory solution to low
emissions and acceptable fuel economy may depend on the advancement of
understanding of the principal gas kinetic and gas dynamic process involved.
Moreover, this may provide a basis for successful development of an
unthrottled prechamber engine.
8-76
-------
APPENDIX A
-------
APPENDIX A
ENVIRONMENTAL PROTECTION AGENCY
PROTOTYPE VEHICLE PERFORMANCE SPECIFICATION
FULL-SIZE PASSENGER CAR
January 3, 1972
Division of Advanced Automotive
Power Systems Development
2929 Plymouth Road
Ann Arbor, Michigan 48105
Approved: —. , ,—
•'Mr. Qfeorge M. Thur
Chief, Power Systems Branch
A-l
-------
ADVANCED AUTOMOTIVE POWER SYSTEMS (AAPS)
PROTOTYPE VEHICLE PERFORMANCE SPECIFICATION'"
January 3, 1972
The AAPS Vehicle performance design specification presented below is intended
to provide:
A common objective for prospective contractors.
Criteria for evaluating proposals and selecting a contractor.
Criteria for evaluating competitive power systems for entering
first generation system hardware.
Advisory criteria to assist the contractor in such areas as
rolling resistance, vehicle air drag etc.
The derived criteria are based on typical characteristics of the class of
passenger automobiles with the largest market volume produced in the U.S.
during the model years 1969 and 1970. It is noted that emissions, volume
and most weight characteristics presented are maximum values while the
performance characteristics are intended as minimum values. Contractors
and prospective contractors who take exceptions must justify these excep-
tions and relate these exceptions to the technical goals presented herein.
Supersedes "Vehicle Design Goals - Six Passenger Automobile"
(Revision C - May 28, 1971)
A-2
-------
CONTENTS
SECTION PAGE
Introduction _ A-4
Vehicle Weight Without Propulsion System 1 A-5
Propulsion System Weight 2 A-5
Vehicle Curb Weight 3 A-5
Vehicle Test Weight 4 A-5
Gross Vehicle Weight 5 A-6
Propulsion System Volume 6 A-6
Air Drag 7 A-6
Rolling Resistance 8 A-7
Propulsion System Emissions 9 A-7
Fuel 10 A-8
Start Up, Acceleration and Grade Velocity •
Performance 11 A-8
Minimum Vehicle Range 12 A-10
Fuel Consumption 13 A-ll
Accessory Power Requirements 14 A-ll
Propulsion System Operating Temperature and
Pressure Range 15 A-ll
Passenger Comfort Requirements 16 A-12
Noise Standards 17 A-12
Operational Life 18 A-12
Reliability and Maintainability 19 A-13
Cost of Ownership 20 A-13
Safety Standards 21 A-13
Test Conditions Table A-1 A-14
A-3
-------
INTRODUCTION
The design of an automobile from a total systems standpoint could be expected
to result in major benefits in cost, safety, and performance. This specifica-
tion is intended as a step along that path, describing a propulsion system that
can be installed into engine compartments as they now exist. Integration of
the vehicle accessories within the propulsion system are highly desirable.
Following the successful demonstration of this development further optimiza-
tion of the propulsion system with the power train, suspension, and vehicle
styling would be possible.
A-4
-------
1. VEHICLE WEIGHT WITHOUT PROPULSION SYSTEM - W
. . o
WQ is the weight of the vehicle excluding the propulsion system. This
weight includes, but is not limited to the frame, body, glass, trim,
suspension, wheels (rims and tires), service brakes, seats, uphol-
stery, sound absorbing materials, insulation, dashboard instruments,
accessory ducting and wiring, accessories, and all other components
not included as part of the propulsion system.
Accessories are defined as driver assistance and passenger convenience
components and subsystems not essential to propulsion system opera-
tion. Included are power steering systems, power brake systems and
passenger compartment heating and air conditioning systems.
W is fixed at 2700 Ibs.
o
2. PROPULSION SYSTEM WEIGHT - W
p
Wp includes the energy storage subsystem (including fuel, containment,
and supply and deliver ducting), power conversion subsystem (including
auxiliaries and control) and power transmitting subsystem (including
transmission and drive train to the driven wheels).
Auxiliaries are defined as components and subsystems essential to the
operation of the power conversion system. Included are electric power
generating subsystems, starting subsystems, exhaust subsystems,
motors fans, flowers, pumps, and fluids.
Lightweight propulsion system are highly desirable, however, the
maximum allowable propulsion system weight, W_m, is 1500 Ibs.
3. VEHICLE CURB WEIGHT - WG
W = W + W .
cop
The maximum allowable vehicle curb weight, Wcm, is 4200 Ibs.
(2700 + 1500 max. = 4200).
4. VEHICLE TEST WEIGHT - Wt
-\yt = W + 400 Ibs. Wt is the vehicle weight at which all accelerative
maneuvers, fuel economy and emissions are to be calculated. (Items
9c, lie, lid, lie, llf, 12, 13).
The maximum allowable test weight, Wtm> is 4600 Ibs.
(2700 + 1500 max. + 400 = 4600).
A-5
-------
5. GROSS VEHICLE WEIGHT - W
o
Wg = Wc + 1100 Ibs. Wg is the gross weight vehicle weight at which
sustained velocity capability at 5 percent and 30% grades is to be cal-
culated (Item llf). The 1100 Ibs. load simulates a full load of pas-
sengers and baggage.
The maximum allowable gross vehicle weight, WgTn, is 5300 Ibs.
(2700 + 1500 max. + 1100 = 5300).
6. PROPULSION SYSTEM VOLUME - V
p
V_ is the volume allotted for all items identified under item 2. The
propulsion system shall be packagable in such a way that the volume
encroachment on either the passenger or luggage compartment does
not exceed the following:
a) The transmission tunnel may not be widened so as to
decrease the selected production vehicle clearance
between the accelerator pedal and the tunnel. The
accelerator pedal may not be relocated.
b) Intrusion of the tunnel into the passenger side of the
vehicle may be increased by a maximum of 1. 5 inches.
c) The tunnel height may be increased by a maximum of
2 inches but without affecting the full fore and aft adjust-
ment of the front seat of the vehicle. The front seat
may not be raised.
The propulsion system shall not violate the vehicle ground clearance
lines as established by the manufacturer of the vehicle used for pro-
pulsion system/vehicle packaging. Additionally, the propulsion system
shall not violate the space allocated for wheel jounce motions and vehicle
steering clearances. Necessary external appearance (styling) changes
will be minor in nature. The propulsion system shall also be packagable
in such a way that the handling characteristics of the vehicle are not
degraded.
7. AIR DRAG
The product of the drag coefficient, C
-------
8. ROLLING RESISTANCE
Rolling resistance, R, is expressed in the equation (R = (W/65
[1 + (1.4 x 10-3V) + (1.2 x 10-5V2)] Ibs. V is the vehicle velocity
in ft/sec. W is the vehicle weight in Ibs.
9. PROPULSION SYSTEM EMISSIONS
The vehicle is to be tested for emissions in accordance with the pro-
cedure in the July 2, 1971 Federal Register and as further described
in the Code of Federal Regulations Title 40 Part 85 for model year
1976 light duty vehicles (CFR 40-85). The vehicle test weight shall
be "W^ and the accessory load as defined in Section 14a. Ambient con-
ditions are 14.7 psia and 85°F. Emission tests will be run with fuel
specified in Section 10.
The Federal emissions standards are:
Hydrocarbons - 0.41 grams /mile maximum
Carbon Monoxide - 3.40 grams/mile maximum
Oxides of Nitrogen* - 0.40 grams/mile maximum
Prototype vehicles are to meet the following emissions goals to allow
for production tolerances and life degradation. Measurement of
emissions is to be taken after the system has operated for 100 hours.
Hydrocarbons - 0. 20 grams/mile maximum
Carbon Monoxide 1.70 grams /mile maximum
Oxides of Nitrogen* - 0.20 grams/mile maximum
'Oxides of nitrogen are to be measured or computed as
Production of smoke, odors, aldehydes, ammonia, particulates or
other undesirable emissions not now specified in the July 2, 1971
Federal Register are undesirable.
A-7
-------
10.
FUEL
Emission tests will use the fuel specified below, however, the power
system shall have the capability of meeting emission levels using com-
mercially available unleaded fuels.
Item
Octane, Research, min.
Pb. (Organic), gm/U.S. gal
Distillation range
I.E.P., °F
10 percent point, °F
50 percent point, °F
90 percent point, °F
E.P. °F (max)
Sulfur, Wt. percent max.
Phosphorous, theory
R.V.P. Ib.
Washed gum (max) mgm/gal
Corrosion (not lower than)
Oxidation stability (not less than)
Hydrocarbon composition
Olefins, percent, max.
Aromatics, percent, max.
Saturates
ASTM
Designation
D1656
D 525
D 86
D1266
D
D
D
D
323
381
130
525
D1319
Specification
91-93
<0.02
100-115
140-150
240-250
330-340
425
0. 10
0.0
5.5-7.5
4.0
IB
240+
30
40
Remainder
For computation purposes the lower heating values of this fuel is to be
assumed as 18500 Btu/lb. The cost to be assumed for system cost
analysis is $0. 3 I/gallon. An A. P. I. gravity of 56. 0 is to be assumed
in all calculations.
11. START UP, ACCELERATION, AND GRADE VELOCITY
PERFORMANCE
a. Startup:
The vehicle must be capable of being tested in accordance with
the procedure outlined in the July 2, 1971 Federal Register with-
out special driver startup/warmup procedures. The accessory
load shall be as defined in Section 14b.
The maximum time from "key on" to reach 65 percent of full
power level is 45 sec. Ambient conditions are 14.7 psia 60°F.
The vehicle is to be soaked at this temperature for a minimum
of 12 hours prior to initiation of start test.
A-8
-------
Powerplant starting procedures in low ambient temperatures shall
be equivalent to or better than the typical automobile spark-
ignition engine. After a 24 hour soak at -20° F and 14.7 psia the
engine shall achieve a self-sustaining idle condition without further
driver input within 25 seconds. No starting aids external to the
normal vehicle system shall be needed at or above -20° F.
b. Idle operation conditions:
The idle creep torque shall not result in level road operation of
the vehicle at a speed in excess of 18 mph in high gear, with the
entire propulsion system at steady state operating temperature
and ambient conditions of 14.7 psia and 85° F. The accessory
load shall be as defined in Section 14a.
c. Acceleration from a standing start:
The minimum distance to be covered in 10. 0 sec. is 440 ft. The
maximum time to reach a velocity of 60 mph is 13.5 sec. Ambient
conditions are 14.7 psia, 85°F. Vehicle weight is W^. and acces-
sory load as defined in Section 14a. Acceleration is on zero
grade and initiated with the engine at the normal idel condition.
d. Acceleration in merging traffic:
The maximum time to accelerate from a constant velocity of
25 mph to a velocity of 70 mph is 15. 0 sec. Ambient conditions
are 14.7 psia, 85° F. Vehicle weight is W^. and accessory load
as defined in Section 14a. Acceleration is on zero grade and
time starts when the accelerator pedal is depressed.
e. Acceleration, DOT High Speed Pass Maneuver:
The maximum time and distance to go from an initial velocity of
50 mph with the front of the automobile (18 foot length assumed)
100 feet behind the back of a 55 foot truck traveling at a constant
50 mph, to a position where the back of the automobile is 100 feet
in front of the front of the 55 foot truck, is 15 sec. and 1400 ft.
The entire maneuver takes place in a traffic lane adjacent to the
lane in which the truck is operated. Vehicle is accelerated until
the maneuver is completed or until a maximum speed of 80 mph
is attained, whichever occurs first. Vehicle acceleration ceases
when a speed of 80 mph is attained, the manuever then being com-
pleted at a constant 80 mph. (This does not imply a design re-
quirement limiting the maximum vehicle speed to 80 mph. ) Time
starts when the accelerator pedal is depressed. Ambient condi-
tions are 14.7 psia, 85° F. Vehicle weight is Wt and accessory
load as defined in Section 14a. Acceleration is on zero grade.
A-9
-------
f. Grade velocity:
The vehicle must be capable of starting from rest on a thirty
percent (30%) grade and ascending the grade at a minimum speed
of 5 mph. The vehicle must be capable of this maneuver in both
the forward and reverse directions at a vehicle weight of Wg,
with the accessory load as defined in Section 14a.
The minimum cruise velocity that can be continuously maintained
on a five percent (5%) grade shall be not less than 65 mph with a
vehicle weight of W and accessory load as defined in Section 14a.
O
The minimum cruise velocity that can be continuously maintained
on a zero percent (0%) grade shall be not less than 85 mph with a
vehicle weight of W^. and with the accessory load as defined in
Item 14a.
Ambient conditions for all grade specifications are 14. 7 psia and
85°F.
Performance degradation attributable to loss of powerplant efficiency
at extreme temperatures shall not exceed ten percent (10%) relative to
the performance values specified at 85° F. This limitation applies to
ambient temperatures from -20°F to 105°F.
The wind velocity is to be less than 10 mph for all acceleration and
grade tests.
12. MINIMUM VEHICLE RANGE
Minimum vehicle range without refueling will be 200 miles (maximum
fuel capacity is 25 U.S. gallons). The minimum range shall be calcu-
lated for, and applied to each of the following modes:
1. Cyclic mode is: The Federal driving cycle which is in accor-
dance with the July 2, 1971 Federal Register.
The range may be calculated for one cycle and
ratioed to 200 miles.
2. Cruise mode is: A constant 70 mph cruise on a zero grade for
200 miles.
The vehicle weight for both modes shall be W, initially and with acces-
sory power levels as specified in Section 14. The ambient conditions
shall be a pressure of 14.7 psia, and a temperature of -20° F (air
conditioner off) and 105° F (air conditioner on).
A-10
-------
13. FUEL CONSUMPTION
Using the fuel specified in Section 10, a "fuel economy" figure shall be
calculated based on 1) miles per gallon and 2) the number of Btu per
mile required to drive the vehicle through the following modes of
operation:
Hours % of Time
1) Federal Driving Cycle 19.84 1750 50
2) Simplified Surburban Route 30.00 1150 33
(equal times at constant
20, 30 and 40 mph speeds).
3) Simplified Country Route 60.00 600 17
(equal times at constant
50, 60 and 70 mph speeds).
Totals 30 3500 100
In all cases the system fuel consumption shall be calculated for a vehicle
weight of W^ initially, and power levels as specified in item 14a. Ambient
conditions are 14.7 psia and 85°F.
It is desirable that the fuel consumption rate at idle operating condition
not exceed 7 Ibs/hour.
14. ACCESSORY POWER REQUIREMENTS
a. Accessory power requirements with the air conditioning in opera-
tion are defined as 15 hp at maximum engine speed and 4 hp at
engine idle speed, with a linear relationship between these two
points.
b. Accessory power requirements without the air conditioning in
operation are defined as 5 hp at maximum engine speed and 2 hp
at engine idle speed, with a linear relationship between these
two points.
15. PROPULSION SYSTEM OPERATING TEMPERATURE AND
PRESSURE RANGE
The propulsion system shall be operable within an expected ambient
temperature range of -40° to 125° F.
The propulsion system shall be operable within an expected environmen-
tal pressure range of 9 psia to 15 psia.
A-ll
-------
16. PASSENGER COMFORT REQUIREMENTS
Heating and air conditioning of the passenger compartment shall be at
a rate equivalent to that provided in the present (1970) standard full size
family car.
Present practice for maximum passenger compartment heating rate is
approximately 30, 000 Btu/hr. For an air conditioning system at 110°F
ambient, 80 ° F and 40% relative humidity air to the evaporator, the rate
is approximately 13,000 Btu/hr.
17. NOISE STANDARDS
»ij
a. Maximum noise test:
The maximum noise generated by the vehicle shall not exceed
77 dbA when measured in accordance with SAE J986a. Note that
the noise level is 77 dbA whereas in the SAE J986 the level is
86 dbA.
b. Low speed noise test:
The maximum noise generated by the vehicle shall not exceed
63 dbA when measured in accordance with SAE J986a except that
a constant vehicle velocity of 30 mph is used on the pass-by.
c. Idle noise test:
The maximum noise generated by the vehicle shall not exceed
62 dbA when measured in accordance with SAE J986a except that
the engine is idling (clutch disengaged or in neutral gear) and the
vehicle is stationary. A 360° survey shall be made, the micro-
phone being 10 feet from the vehicle perimeter.
18. OPERATIONAL LIFE
The design lifetime of the propulsion system in normal operation will
be 3500 hours minimum.
Terminationof the operational life of an engine shall be determined by
structural or functional failure. Functional failure is defined as power
degradation exceeding 25 percent of maximum output of the rear wheels.
The air conditioner will not be in operation during noise tests.
A-12
-------
19. RELIABILITY AND MAINTAINABILITY
The reliability and maintainability of the vehicle shall equal or exceed
that of the spark-ignition automobile. The mean-time-between-failure
should be maximized to reduce the number of unscheduled service trips.
No failure modes shall present a serious safety hazard during vehicle
operation and servicing. Failure propagation should be minimized.
The power plant should be designed for ease of maintenance and repairs
to minimize costs, maintenance personnel education, and downtime.
20. COST OF OWNERSHIP
The initial cost and net cost of ownership of the vehicle shall be minimized
for ten years and 105, 000 miles of operation.
21. SAFETY STANDARDS
The vehicle shall comply with all Department of Transportation Federal
Motor Vehicle Safety Standards in force when the selected test vehicle
was manufactured.
A-13
-------
TABLE A-l (Excerpt)
TEST CONDITIONS
(ENGINE DYNAMOMETER, CHASSIS DYNAMOMETER, ROAD)
Section
9. EMISSIONS
11. PERFORMANCE
a. Start Up
b. Idle
c. Accel (0-60)
d. Accel (25-70)
e. Accel (50-80)
f. Grade (30%)
(5%)
(0%)
12. VEHICLE RANGE
13. FUEL CONSUM.
NOTES: Emission tests
Road test wind
Minimum values
Maximum values
Performance Requirements
HC 0. 20 grams per mile, ....
CO 1.70 grams per mile.
N©2 0. 20 grams per mile
65% power in 45 sec
Driver Assistance 25 sec
**
Creep 18 mph
13. 5 sec' " to 60 MPH
440 ftv in 10.0 sec
15. 0 sec """ from 25 to 70 MPH
15.0 sec""'1 and 1400, ft
From 50 to 80 MPH^
0 to 5 MPH*
65 MPH*
85 MPH*
200 MI*
1. During FCE
2. At 70 MPH
MPG During FDC
MPG at 20, 30, and 40 MPH
MPG at 50, 60, and 70 MPH
Accessory
Power
14a
14a
14a
14b
14b
14a
14a
14a
14a
14a
14a
14a
14a
14b and 14a
14b and 14a
14a
14a
14a
Weight
W
W
wj
W
wt
w
wt
wt
w
wg
t
w
wl
w
w
w
Temperature
85°F
85°F
85°F
60°F
-20°F
85°F
85°F
85°F
85°F
85°F
85°F
85°F
85°F
-20°F and 105°F
-20°F and 105°F
85°F
85°F
85°F
Pressure
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
14.7
will be run with fuel specified in Section 10.
conditions shall not exceed 10 MPH in any direction.
-------
APPENDIX B
-------
APPENDIX B
ATTACHMENT
ENVIRONMENTAL PROTECTION AGENCY
PRELIMINARY
PROTOTYPE VEHICLE SPECIFICATION
COMPACT PASSENGER CAR
FEBRUARY 1974
Alternative Automotive Power
Systems Division
2929 Plymouth Road
Ann Arbor, Michigan 48105
B-l
-------
ALTERNATIVE AUTOMOTIVE POWER SYSTEMS DIVISION (AAPSD)
PROTOTYPE VEHICLE SPECIFICATION
COMPACT PASSENGER CAR
This specification defines emissions, fuel economy, noise and performance
requirements for automobiles intended to be powered by other than spark
ignition engines. Emissions, volume and most weight characteristics pre-
sented are maximum values. Performance characteristics are intended as
minimum values.
Significant changes over previous specifications are in the vehicle size and
performance requirements.
B-2
-------
CONTENTS
SECTION PAGE
Vehicle Weight Without Propulsion System 1 B-4
Propulsion System Weight 2 B-4
Vehicle Curb Weight 3 B-4
Vehicle Test Weight 4 B-4
Gross Vehicle Weight 5 B-5
Propulsion System Volume 6 B-5
Air Drag 7 B-5
Rolling Resistance 8 B-6
Propulsion System Emissions 9 B-6
Fuel 10 B-7
Start Up, Acceleration and Grade
Velocity Performance 11 B-8
Minimum Vehicle Range 12 B-9
Fuel Economy 13 B-9
Accessory Power Requirements 14 B-ll
Propulsion System Operating Temperature
and Pressure Range 15 B-ll
Passenger Comfort Requirements 16 B-ll
Noise Standards 17 B-12
Operational Life 18 B-12
Reliability and Maintainability 19 B-12
Cost of Ownership 20 B-12
Safety Standards 21 B-13
Vehicle Handling 22 B-13
B-3
-------
1. VEHICLE WEIGHT WITHOUT PROPULSION SYSTEM - W
o
Wo is the weight of the vehicle excluding the propulsion system. This
weight includes, but is not limited to, the frame, body, glass, trim,
suspension, wheels (rims and tires), service brakes, seats, upholstery,
sound absorbing materials, insulation, dashboard instruments, acces-
sory ducting and wiring, accessories, and all other components not
included as part of the propulsion system.
Accessories are defined as driver assistance and passenger convenience
components and subsystems not essential to propulsion system opera-
tion. Included are power steering systems, power brake systems and
passenger compartment heating and air conditioning systems.
The maximum allowable weight W is 2000 Ibs.
2. PROPULSION SYSTEM WEIGHT - W
p
Wp includes the fuel storage subsystem (including fuel, containment,
and supply and deliver ducting), power conversion subsystem (including
auxiliaries and controls) and power transmitting subsystem (including
transmission and drive train to the driven wheels).
Auxiliaries are defined as components and subsystems essential to the
operation of the propulsion system. Included are electric power
generating subsystem, starting subsystem, exhaust subsystem, motors,
fans, blowers, pumps, auxiliary drives, and fluids.
The maximum allowable propulsion system weight, W , is 1000 Ibs.
3. VEHICLE CURB WEIGHT - WG
W = W + W .
cop
The maximum allowable vehicle curb weight, W . is 3000 Ibs.
6 ' cm
4. VEHICLE TEST WEIGHT - Wt
Wt = Wc + 400 Ibs. Wj. is the vehicle weight at which all accelerative
maneuvers, fuel economy and emissions are to be calculated. (Items 9,
lie, lid, llf, 12, 13.)
The maximum allowable test weight, Wtm, is 3400 Ibs. For chassis
dynamometer testing, maximum inertia weight is 3500 Ibs.
B-4
-------
5. GROSS VEHICLE WEIGHT - W
g
Wg = Wc + 700 Ibs. Wg is the gross vehicle weight at which the DOT
pass maneuver (lie) and sustained velocity capability at 30% grades is
to be calculated. (Item llf. ) The 700 Ib load simulates a full load
of four passengers and baggage.
The maximum allowable gross vehicle weight, W , is 3700 Ib .
&
6. PROPULSION SYSTEM VOLUME - V
p
Vp is the volume allotted for all items identified under item 2. The
propulsion system shall be packagable in such a way that the volume
encroachment on either the passenger or luggage compartment does
not exceed the following:
a) The transmission tunnel may not be widened so as to
decrease the selected production vehicle clearance
between the accelerator pedal and the tunnel. The
accelerator pedal may not be relocated.
b) Intrusion of the tunnel into the passenger side of the
vehicle may be increased by a maximum of 1. 5 inches.
c) The tunnel height may be increased by a maximum of
2 inches but without affecting the full fore and aft
adjustment of the front seat of the vehicle. The front
seat may not be raised.
The propulsion system shall not violate the vehicle ground clearance
lines as established by the manufacturer of the vehicle used for pro-
pulsion system/vehicle packaging. Additionally, the propulsion system
shall not violate the space allocated for wheel jounce motions and vehicle
steering clearances. Necessary external appearance (styling) changes
will be minor in nature. The propulsion system shall also be packag-
able in such a way that the handling characteristics of the vehicle are
not degraded.
7. AIR DRAG
The product of the drag coefficient, Cd, and the frontal area, Af, to
be used in air drag calculations shall be 10 ftz. The air density used
in computations shall correspond to the applicable ambient air
temperature.
B-5
-------
8. ROLLING RESISTANCE
Rolling resistance, R, is expressed in the equation
R = (W/65) [1 + (1.4 x 10-3V) + (1.2 x 1Q-5V2)] Ibs. V is the vehicle
velocity in ft/sec. W is the vehicle weight in Ibs.
9. PROPULSION SYSTEM EMISSIONS
a. Exhaust Emissions
The vehicle is to be tested for emissions in accordance with the pro-
cedure in the November 15, 1972 Federal Register for model year
1975 light duty vehicles. The vehicle test weight shall be as specified
in the Federal Register for the actual curb weight W^.. The accessory
load shall be as defined in Section 14 without air conditioning in opera-
tion. Ambient conditions are 14.7 psia and 85° F. Emissions tests will
be run with fuel specified in Section 10.
The applicable Federal emissions standards are:
Hydrocarbons - 0.41 grams/mile maximum
Carbon Monoxide - 3.40 grams/mile maximum
Oxides of Nitrogen''1 - 0.40 grams/mile maximum
Prototype Vehicles shall meet the following emissions goals when tested
as specified above. Measurement of emissions is to be taken after the
system has accumulated 100 hours.
Hydrocarbons - 0.20 grams/mile maximum
Carbon Monoxide - 1.70 grams/mile maximum
Oxides of Nitrogen'1' - 0. 20 grams/mile maximum
*
Oxides of nitrogen are to be measured or computed as NO?.
Production of smoke, odors, aldehydes, ammonia, particulates or
other emissions not now specified in the November 15, 1972 Federal
Register are undesirable. Tests for smoke shall utilize instrumenta-
tion in accordance with paragraph 85. 125 of the Federal Register
November 10, 1970. Tests for odor shall be by the Quality Intensity
(Turk) method.
b. Fuel Evaporative Emissions
Fuel Evaporative Emissions shall be measured in accordance with the
procedure in the November 15, 1972 Federal Register. The Fuel
Evaporative Emission standard is 2.0 grams per test.
B-6
-------
10. FUEL
a.
Test Fuel
Emission tests will use the fuel specified below, however, the power
system shall have the capability of meeting the applicable Federal
emission standards above using other commercially available fuels as
stated in paragraph lOb.
'F
'F
'F
Item
Octane, Research, min.
Pb (Organic) gm/U.S. gal
Distillation range
I.E.P., °F
10 percent point,
50 percent point,
90 percent point,
E.P. °F (max)
Sulfur, Wt. percent max.
Phosphorous, theory
R.V.P. Ib.
Washed gum (max mgm/100 ml)
Corrosion (not lower than)
Oxidation stability (not less than)
Hydrocarbon Composition
Olefins, percent, max.
Aromatics, percent, max.
Saturates
Nitrogen, Wt % max. *v
ASTM
Designation
D2699
D 524
D 86-67
D1266
D
D
D
D
323
381
130
525
D1319
Specification
91-93
<0. 02
100-115
140-150
240-250
330-340
425
0. 10
0.0
5.5-7.5
4.0
IB
240+
30
40
Remainder
0.005
r ^Nitrogen content shall include chemically bound plus additiye
introduced. Measurement shall be by Kjeldahl method.
For computation purposes the lower heating value of this fuel is to be
assumed as 18500Btu/lb. An A.P.I, gravity of 56.0 is to be assumed
in all calculations except where actual measurements are available.
b.
Alternative Fuels
In order to accommodate future requirements for multi-fuel capabilities,
new designs should consider presently available transportation fuels
such as Jp5, Jp4, Kerosene, unleaded gasoline, Diesel f 1 and Diesel
#2. Where using fuel other than unleaded gasoline, minor modifications
to the fuel and combustion system are acceptable.
B-7
-------
11. START UP, ACCELERATION, AND GRADE VELOCITY
PERFORMANCE
The wind velocity is to be less than 10 mph for all acceleration and
grade tests.
a. Startup
The vehicle must be capable of being tested in accordance with the pro-
cedure outlined in the November 15, 1972 Federal Register without
special driver startup/warmup procedures. The accessory load shall
be as defined in Section 14, without air conditioning in operation.
The start up performance characteristics of the engine shall be demon-
strated by achieving a self sustaining idle condition within 30 seconds
from "key on", followed by a 10 second period of increasing power
equivalent to an acceleration rate, from a standing start, of 3 miles
per hour per second for a vehicle weight of W^ on a zero grade.
Ambient pressure shall be 14.7 psia. Ambient temperature shall be
between 68 and 85° F.
Powerplant starting procedures at low ambient temperatures shall be
equivalent to or better than the typical pre-control automobile spark-
ignition engine. No starting aids external to the normal vehicle system
shall be needed after a 24 hour soak at -20°F.
b_. Idle Operation Conditions:
The idle creep torque shall not result in level road operation of the
vehicle at a speed in excess of 15 mph in high gear, with the entire
propulsion system at steady state operating temperature and ambient
conditions of 14.7 psia and 85° F. The accessory load shall be as
defined in Section 14, without air conditioning in operation.
c. Acceleration from a Standing Start:
The maximum distance to be covered in 10.0 sec. is 440 ft. The maxi-
mum time to reach a velocity of 60 mph is 17. 5 sec. Ambient conditions
are 14.7 psia, 85° F. Vehicle weight is W^ and accessory load as de-
fined in Section 14 without air conditioning in operation. Acceleration
is on zero grade and time starts when the accelerator pedal is depressed.
el. Acceleration in Merging Traffic:
The maximum time to accelerate from a constant velocity of 25 mph
to a velocity of 70 mph is 20.0 sec. Ambient conditions are 14.7 psia,
85° F. Vehicle weight is W and accessory load as defined in Section 14
B-8
-------
without air conditioning in operation. Acceleration is on zero grade
and time starts when the accelerator pedal is depressed.
e. Acceleration, DOT High Speed Pass Maneuver:
The maximum time to complete the DOT High Speed Pass Maneuver is
17 seconds. The entire maneuver takes place in a traffic lane adjacent
to the lane in which the truck is operated. Vehicle is accelerated until
the maneuver is completed or until a maximum speed of 80 mph is
attained, whichever occurs first. Vehicle acceleration ceases when a
speed of 80 mph is attained, the maneuver then being completed at a
constant 80 mph. Time starts when the accelerator pedal is depressed.
Ambient conditions are 14.7 psia, 85°F. Vehicle weight is Wg and
accessory load as defined in Section 14 without air conditioning in opera-
tion. Acceleration is on zero grade.
f. _ Grade Velocity;
The vehicle must be capable of starting from rest on a thirty percent
(30%) grade and ascending the grade at a minimum speed of 5 mph.
The vehicle must be capable of this maneuver in both the forward and
reverse directions at a vehicle weight of Wg, with the accessory load
as defined in Section 14, without air conditioning in operation.
12. MINIMUM VEHICLE RANGE
Minimum vehicle range without refueling will be 300 miles for a constant
50 mph cruise on a zero grade.
The vehicle weight shall be Wg initially and with accessory power levels
as specified in Section 14. The ambient conditions shall be a pressure
of 14.7 psia, and a temperature of -20°F (air conditioner off) and
105°F (air conditioner on). Maximum fuel capacity shall be 15 U.S.
gallons.
13. FUEL ECONOMY
a. Fuel Economy of Current Vehicles
The average fuel economy over the Federal Driving Cycle for 1400 tests
on 1973 model year vehicles was reported in SAE paper 730790. Cor-
recting that analysis for the 1975 test procedure gives the following
equation for fuel economy as a function of weight:
MPG(avg) =^f^- 1.34 (14.9mpgfor Wt = 3400 Ib)
B-9
-------
b. Fuel Economy Goal
The goal for first generation prototype hardware is to equal 130% of the
fuel economy shown by the equation. (1.3 x 14.9 = 19.4 mpg for
wt = 3400 Ib).
In addition to the goal stated above, fuel economy shall be measured
for the component parts of the following driving profile, and the equiv-
alent miles per gallon calculated for the overall model.
Mode
Federal
Driving Cycle
30 mph
40 mph
50 mph
60 mph
70 mph
Total
Mode Usage
465 23-minute runs
1 Cold start each run
50 12-minute runs
1 Hot start each run
125 12-minute runs
1 Hot start each run
150 12-minute runs
1 Hot start each run
175 12-minute runs
1 Hot start each run
114 12-minute runs
1 Hot start each run
465 Cold starts,
614 Hot starts
Miles
Per Year
3500
300
1000
1500
2100
1600
10, 000 mi.
Hours
Per Year
177
10
25
30
35
23
300 hr.
The goal for idle fuel consumption will not be greater than 5 pounds
per hour.
In all cases, fuel economy shall be determined including suitable acces-
sory loads in accordance with Section 14. A separate determination
shall be made with and without air conditioning. The goals stated above
shall be met without air conditioning. Ambient conditions are 14.7 psia
and 85° F. Cold start means the engine has been soaked at the ambient
condition for at least 12 hours. Hot start means the engine is started
with no greater than 15 minutes elapsed from the prior complete shut
down.
For prototype hardware, a complete map of BSFC shall be obtained,
with fuel consumption referenced to the horsepower at the transmission
input coupling, and all engine auxiliaries deriving power from the engine
output torque. (No Test Support Equipment.)
B-10
-------
14. ACCESSORY POWER REQUIREMENTS
In all cases accessory power consumption shall be minimized.
Consideration shall be given to accessory power subsystems which
are de-coupled from the torque-speed-power characteristics of the
engine. In the absence of other experimental data, the following
equation shall be used to determine accessory power. Without air
conditioning in operation:
hp = 1. 5 + (. 333 x 10"3) (rpm )
With air conditioning in operation:
hp = 3.0 + (1.83 x 10"3) ( rpm)
where hp = accessory power
rpm = shaft speed of the accessory drive shaft
15. PROPULSION SYSTEM OPERATING TEMPERATURE
AND PRESSURE RANGE
The propulsion system shall be operable within an expected ambient
temperature range of -40° to 125°F.
The propulsion system shall be operable within an expected ambient
pressure range of 9 psia to 15 psia.
Performance degradation attributable to loss of powerplant efficiency
at extreme temperature shall not exceed ten percent (10%) relative
to the performance values specified at 85° F. This limitation applies
to ambient temperatures from -20° F to 105° F.
16. PASSENGER COMFORT REQUIREMENTS
Heating and air conditioning of the passenger compartment shall be at
a rate equivalent to that provided in the present compact size passenger
car.
Present practice for maximum passenger compartment heating rate is
approximately 20,000 Btu/hr. For an air conditioning system at 110°F
ambient, 80° F and 40% relative humidity air to the evaporator, the rate
is approximately 10, 000 Btu/hr.
B-il
-------
17. NOISE STANDARDS
a. Maximum Noise
The maximum noise generated by the vehicle shall not exceed 77 dbA
when measured in accordance with SAE J986a. Note that the noise
level is 77 dbA whereas in the SAE J986 the level is 86 dbA.
b. Low Speed Noise
The maximum noise generated by the vehicle shall not exceed 63 dbA
when measured in accordance with SAE J986a except that a constant
vehicle velocity of 30 mph is used on the pass-by.
c. Idle Noise
The maximum noise generated by the vehicle shall not exceed 62 dbA
when measured in accordance with SAE J986a except that the engine is
idling (clutch disengaged or in neutral gear) and the vehicle is sta-
tionary. A 360° survey shall be made, the microphone being 10 feet
from the vehicle perimeter.
*
The air conditioner will not be in operation during noise tests.
18. OPERATIONAL LIFE
The design lifetime of the propulsion system in normal operation will
be 3500 hours minimum.
Termination of the operational life of an engine shall be determined by
structural or functional failure. Functional failure is defined as power
degradation exceeding 25 percent of maximum output of the rear wheels.
19. RELIABILITY AND MAINTAINABILITY
The reliability and maintainability of the vehicle shall equal or exceed
that of the spark-ignition engine automobile. The mean-time-between-
failure should be maximized to reduce the number of unscheduled
service trips. No failure modes shall present a serious safety hazard
during vehicle operation and servicing. Failure propagation should be
minimized. The power plant should be designed for ease of maintenance
and repairs to minimize costs, maintenance personnel education, and
downtime.
20. COST OF OWNERSHIP
The initial cost and net cost of ownership of the vehicle shall be mini-
mized for ten years and 105, 000 miles of operation.
B-12
-------
21. SAFETY STANDARDS
The vehicle shall comply with all Department of Transportation Federal
Motor Vehicle Safety Standards in force when the selected test vehicle
was manufactured.
22. VEHICLE HANDLING
The use of an alternative engine shall not significantly change the
handling characteristics ("feel") of the vehicle. To demonstrate this,
the vehicle handling characteristics of alternative engine powered
vehicles will be compared to conventionally powered vehicles with as
close to the same overall configuration as possible. The handling per-
formance goals shall be:
a. Low Speed Maneuverability (Slalom)
The maximum safe speed at which the vehicle can maneuver through
pylons will be determined. Pylon spacing will be 48 feet, with 12 feet
wide lanes. Maximum test speed will be 25 mph.
b. Braking Distance
The braking distance will be measured from 30 mph to stop within a
12 foot wide lane.
c. Cornering
The maneuverability of the vehicle into and around both 50 and 100 foot
diameter circles will be determined. Maximum test speed on the
50 foot circle will be 15 mph; and maximum test speed on the 100 foot
circle will be 22 mph.
d. Lane Change (Swerve)
The maximum safe speed for changing from one 12 foot wide lane to
another 12 foot wide lane and back over a total distance of 240 feet
will be determined. The maximum test speed will be 50 mph.
e. Smoothness (Driveability)
The smoothness of the engine-vehicle operation at idle and between
25 and 35 mph will be determined, including the relative idle rough-
ness, hesitation, tip-in, stumble, surge, detonaton, backfire, stall
and after-run (dieseling) as applicable.
B-13
-------
ABBREVIATIONS
-------
ABBREVIATIONS
AAPS
ac
Ah
Ah/cm2
ALRC
AMC
AMMRC
APCO
API
APL
ARB
ARPA
ASME
ASTM
BDC
bhp
BSFC
BMEP
Btu
CoHr
£. b
CCCP
CECO
CFR
Alternative Automotive Power Systems
alternating current
ampere hours
ampere hours per square centimeter
Aerojet Liquid Rocket Company
Army Materiel Command
Army Materials and Mechanics Research Center
Air Pollution Control Office (EPA), formerly NAPCA
a method (American Petroleum Institute) of expressing
specific gravity of hydrocarbons:
A pi _ _ 141.5 - __ j 3 1 5
A±^ - specific gravity 60°F/60°F
Applied Physics Laboratory. John Hopkins University
Air Resources Board (California)
Advanced Research Projects Agency (DDR&E)
American Society of Mechanical Engineers
American Society for Testing Materials
bottom dead center
brake horse power
brake specific fuel consumption
brake mean effective pressure
British thermal unit
chemical formula for ethanol (same as ethyl alcohol)
California Clean Car Project
Chandler Evans Control Systems
Cooperative Fuel Research (Council)
chemical formula for methane
Ab-1
-------
CH3OH
CNG
CO
co2
CI
CID
CVCC
CVS
dB(A)
dc
DCF
DOT
EFC
EFE
EGR
EMT
EPA
EVC
FCP
FDC
FSTP
FTP
g/mi
gvm
H,
chemical formula for methyl alcohol (methanol)
compressed natural gas
chemical formula for carbon monoxide
chemical formula for carbon dioxide
compression ignition
cubic inches of displacement
Compound Vortex Controlled Combustion (Honda)
constant volume sampling
decibel referred to one ampere
direct current
Discounted Cash Flow, a. method of computing return on
investment
Department of Transportation
electronic fuel control
early fuel injection
exhaust gas recirculation
electromechanical transmission
Environmental Protection Agency (formerly HEW)
external vaporizing combustion
Ford Combustion Process
Federal Driving Cycle(test for determining exhaust emissions)
Federal Smoke Test Procedure
Federal Test Procedure
grams per mile (also gm/mi)
grams per vehicle mile
chemical formula for hydrogen
Ab-2
-------
HC
HEW
hp
hp-hr
ICE
ihp
IMEP
ISFC
JIC
KKM-125
kW
LAS
Ib
Ib/ihp-hr
LMC
LPG
MCP
MON
mpg
MTI
NAPCA
NH,
N
HA
L* *•
hydrocarbon
Department of Health, Education, and Welfare (predecessor to
EPA in studying the control of automotive emissions)
horsepower
horsepower per hour
internal combustion engine
indicated horsepower
indicated mean effective pressure
indicated specific fuel consumption
jet-induced circulation
Kreiskolbenmotor, orbiting piston engine, 125 cubic centimeters
in volume
kilowatt
lithium- alumina- silicate
pound
pounds per indicated horsepower hour
Lear Motors Corporation
liquified petroleum gas
Mitsubishi Combustion Process
motor octane number
miles per gallon
Mechanical Technology, Inc.
National Air Pollution Control Administration (see APCO)
chemical formula for ammonia
chemical formula for nitrogen
chemical formula for hydrozene
Ab-3
-------
NO
x
°2
PCU
PWM
PROCO
psi
psia
psig
PTX
rms
RON
rpm
SCE
SCFM
SCR
SCRTD
SES
SFC
SI
Si3N4
SNG
SPS
SW RI
TACOM
chemical formula for oxides of nitrogen
chemical formula for oxygen
power control unit
pulse-width modulation
Programmed Combustion (Ford)
pound per square inch
pound per square inch, absolute
pound per square inch, gage
Engelhard PTX, oxidation catalyst
root mean square
research octane number
revolutions per minute
stratified charge engine
standard cubic feet per minute flow of gas, measured
at 70°F and 14.696 psia
silicon controlled rectifier
Southern California Rapid Transit District
Scientifc Energy Systems
specific fuel consumption
spark ignition
silicon nitride
synthetic natural gas
Steam Power System
Southwest Research Institute
(U.S. Army) Tank Automotive Command
Ab-4
-------
TCCS
TC-25
TCP
TDC
TECO
UACRL
USATACOM
V
V
P
W
Wh
Wh/lb
WOT
Texaco Controlled Combustion System
transit coach
Texaco Combustion Process
top dead center
Thermo Electron Company
United Aircraft Corp. Research Laboratory
(see TACOM)
volt
propulsion system volume (see Appendix A)
watt
watt hour
watt hours per pound
wide-open throttle
Ab-5
-------
GLOSSARY
-------
GLOSSARY
aldehydes
Anhydrous
aspiration (British)
autoignition
base metal
brake mean effective
pressure (BMEP)
brake horsepower
brake specific fuel
consumption
Bray ton cycle
catalyst
catalytic converter
cetane number
displacement
an odorous, oxygenated hydrocarbon classified as
a pollutant
without water
process by which engine takes in air fuel mixture
(breathing)
spontaneous ignition
a metal or alloy (as zinc, lead, or brass) of com-
paratively low value and relatively inferior in cer-
tain properties (as resistance to corrosion)--
opposed to noble metal
the effective constant pressure required to be
exerted during each stroke to produce the power
delivered by the engine
the net power available at the output shaft of the
engine
the fuel flow rate required per unit of net horse-
power output of the engine
a thermodynamic cycle, consisting of gas compres-
sion heating and expansion stages without phase
changes of gas; the operating basis of a gas turbine
engine
a substance that increases the rate of a chemical
reaction without itself undergoing a permanent
chemical change
as applied to automotive heat engines, a device
relying on a catalyst in the engine exhaust with
the objective of lessening pollutant emissions
to the atmosphere
indicates the ignition delay of a fuel in a compres-
sion ignition (diesel) engine
the volume swept out by the pistons or other power
elements
Gl-1
-------
endothermic reaction
equivalence ratio
exothermic reaction
flash point
gross or higher
heating value
hypergolic ignition
investment casting
kinematic viscosity
naphtha
natural gas
neat
net or lower
heating value
octane number
Rankine cycle engine
recuperator
chemical reaction in which heat is absorbed
the actual fuel air ratio to stoichiometric fuel air
ratio
a chemical reaction in which heat is evolved
the temperature at which the bulk liquid can be
ignited by an open flame
the energy released by a fuel when it is completely
burned in such a way that the vapor of combustion
is condensed to a liquid
spontaneous ignition upon contact of the fuel and
oxidizer
a process used in metal casting that consists of
making a wax model, coating it with a refractory
to form a mold, heating until the wax melts, and
then pouring metal into the space left vacant
viscosity divided by density
a light oil product of petroleum having properties
intermediate between gasoline and kerosene
primarily a gaseous fuel consisting of methane
a pure material without additives
the energy released by a fuel when it is burned
completely without condensing the water vapor
of combustion
a measure of the resistance of a fuel to detonate in
a spark-ignition, reciprocating engine (Otto cycle
engine)
an external combustion engine in which a high-
pressure working fluid is converted to vapor by
the heat from combustion gases and then is
expanded in a piston or turbine expander to produce
work.
a fixed device used to improve gas turbine engine
efficiency by providing for heat exchange between
turbine exhaust gases and compressor outlet air
Gl-Z
-------
regenerator
regenerative engine
sink
Stirling cycle engine
stoichiometric
syncrude
thermal efficiency
thermal reforming
torque
a rotating device used to improve gas turbine engine
efficiency by providing for heat exchange between
turbine exhaust gases and compressor outlet air
incorporates a regenerator or recuperator to
improve engine efficiency
a body or substance used for the disposal of a
fluid or heat in the course of a hydrodynamic or a
thermodynamic process (as the condenser of a
steam engine)
an external combustion, closed cycle, piston-type
device that uses a gaseous internal working fluid,
usually hydrogen or helium. Cyclical heating and
cooling varies the pressure of the fluid within a
closed volume, the pressure variations being
transmitted to a piston, thereby developing output
power.
refers to conditions in a combustion process where-
in the fuel is completely burned with no excess
oxygen remaining
a synthetic crude oil made from oil shale or coal
the ratio of theoretical work done by an engine to
the mechanical equivalent of heat supplied in the
fuel
a process using heat (but no catalyst) to effect
molecular rearrangement of a hydrocarbon fuel
(example: low-octane naphtha into gasoline of
higher antiknock quality)
any turning movement exerted by a force acting
tangentially at a distance from the axis of rota-
tion, the product of a force and the distance of
its line of action from the axis
Gl-3
-------
REFERENCES
-------
REFERENCES
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Committee on Public Works, United States Senate, by the Environ-
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R-l
-------
3-4 W. R. Wade, et al. , "Low Emission Combustion for the Regenerative
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and Assessment, " ASME paper 73-GT-12, presented at the Gas
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3-6 T. F. Nagey, et al. , "The Low Emission Gas Turbine Passenger
Car - What Does the Future Hold, " ASME paper 73-GT-49, pre-
sented at the Gas Turbine Conference and Product Show,
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3-7 D. J. White, et al. , "Low Emission Variable Area Combustor for
Vehicular Gas Turbine, " ASME paper 73-GT-19, presented at the
Gas Turbine Conference and Product Show (April 1973).
3-8 C. F. McDonald, "Gas Turbine Recuperator Technology Advance-
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Conference and Product Show (March 1972).
3-9 Corning Glass Works, Cercor Glass - Ceramic Rotary Heat
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3-12 Economic Impact of Mass Production of Alternative Low Emissions
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3-14 Final Report - Automotive Gas Turbine Economic Analysis, Williams
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3-15 Automobile Gas Turbine - Optimum Cycle Selection Study, General
Electric Company, Space Division, Cincinnati, Ohio, Report GESP-
725 FS (June 1972).
R-2
-------
3-16 G. A. Amann, W. R. Wade, and M. K. Yu, "Come Factors Affecting
Gas Turbine Passenger Car Emission," SAE paper 720237
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3-17 Briefing documentation, EPA Advanced Automotive Power Systems
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3-18 "GM Goes All Out for Fuel Economy, " Detroit News (February 10,
1974). '
3-19 L. D. Verrelli and C. J. Andary, Exhaust Emission Analysis of the
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Michigan, Report 72-27 (May 1972).
3-20 Li. D. Verrelli, Exhaust Emission Analysis of the Williams Research
Gas Turbine Volkswagen, EPA, Ann Arbor, Michigan, Report 71-30
(May 1971).
3-21 Final Report, Automotive Gas Turbine Optimum Configuration Study,
United Aircraft Research Laboratories, UARL Report L-971249-7
(May 1972)
3-22 History of Chrysler Corporation Gas Turbine Vehicles; March 1954 -
June 1966, Chrysler Corporation (August 1966).
3-23 D. J. White, P- B. Roberts, and W. A. Compton, "Solar JIC-B Com-
bustor Development for Baseline Gas Turbine Engine, " paper presented
at EPA Division of Advanced Automotive Power Systems Development
Contractors Coordination Meeting, Ann Arbor, Michigan (May 14, 1974).
3-24 O. I. Ford, "Low Emission Automotive Gas Turbine Combustor
Development Progress, " Technical Paper for Fourth Gas Turbine
Contractors Coordination Meeting, Ann Arbor, Michigan (December
12-15, 1973).
3-25 Low NO Combustor Program, AiResearch Manufacturing Company,
AT 6154^R (December 13, 1972).
3-26 "Development of Low Emission Porous Plate Combustor for Automo-
tive Gas Turbine and Rankine Cycle Engine, "Final Report, General
Electric Company (July 1973).
3-27 A. F. McLean, et al. , Brittle Materials Design, High Temperature
Gas Turbine, AMMRC CTR 72-19 (September 1972).
3-28 P. M. Ardans and D. W. Stephenson, "An Analytical Method for Esti-
mating the Performance of a Gas Turbine Engine with Water-Methanol
Injection, "SAE paper 700208 (1970).
R-3
-------
3-29 Low NO Emission Combustor for Automobile Gas Turbine Engines,
United Aircraft of Canada, Final Report No. ER 700, EPA Contract
68-04-0015 (February 1973).
3-30 1973 Quarterly Program Reports to EPA under AAPS Program,
Chrysler Corp., Contract No. 68-01-0459.
3-31 Chrysler 5th Quarterly Program Report to EPA (January 31, 1974).
3-32 General Motors letter to The Aerospace Corporation, dated July 1,
1974.
4-1 Automotive Rankine Cycle Contractors Coordination Meeting, Sum-
mary Report. AAPS Div. EPA, Ann Arbor, Michigan (September
1971).
4-2 Automotive Power Systems Contractors Coordination Meeting, Sum-
mary Report. AAPS Div. EPA, Ann Arbor, Michigan (January 1972).
4-3 Automotive Power Systems Contractors Coordination Meeting, Third
Summary Report. AAPS Div. EPA, Ann Arbor, Michigan (June 1972).
4-4 Automotive Power Systems Contractors Coordination Meeting, Fourth
Summary Report. AAPS Div. EPA, Ann Arbor, Michigan (December
1972).
4-5 Advanced Automotive Power Systems Contractors Coordination Meet-
ing, Fifth Summary Report. AAPS Div. EPA, Ann Arbor, Michigan
(June 1973).
4-6 AiResearch Manufacturing Co. , "Compact Condensers for Rankine
Cycle Engines, " presentation, Automotive Rankine Cycle Contractors
Coordination Meeting, AAPS Div. EPA, Ann Arbor, Michigan (Sept-
ember 1971).
4-7 Scientific Energy Systems Corp. , presentation, Rankine Cycle
Contractors Coordination Meeting, AAPS Div. EPA, Ann Arbor,
Michigan (October 1973).
4-8 Thermo Electron Corp. , "Rankine Cycle Power System with Organic
Based Working Fluids and Reciprocating Expander for Passenger
Vehicles, " presentation, Automotive Rankine Cycle Contractors
Coordination Meeting, AAPS Div. EPA, Ann Arbor, Michigan
(October 1973).
R-4
-------
4-9 M. Wenstrom and R. A. Renner, "The California Clean Car Project/'
presentation, Automotive Rankine Cycle Contractors Coordination
Meeting, AAPS Div. EPA, Ann Arbor, Michigan (October 1973).
4-10 O. B. Platell, "Some Contributions to Design of Critical Components
for Automotive Rankine Systems, "SAAB-SCANIA Aerospace Div. ,
presentation, Automotive Rankine Cycle Contractors Coordination
Meeting, AAPS Div. EPA, Ann Arbor, Michigan (October 1973).
4-11 R. A. Renner, "The California Steam Bus Project: Technical Evalu-
ation, " International Research and Technology Corp. , Washington,
D.C. Report IRT-301-R (January 1973).
4-12 AAPS Div. EPA, Private Communication to Aerospace Corporation
(January 1974).
4-13 S. S. Miner, "Developments in Automotive System Power Plants,"
SAE Paper 690043 (January 1969).
5-1 N. D. Postma, et al. , "The Stirling Engine for Passenger Car
Application, " SAE Paper 730648 presented at the Combined Com-
mercial Vehicle Engineering and Operations, Chicago (June 1973).
5-2 R. J. Meijer, "The Philips Stirling Engine, " De Ingenieur (1969).
5-3 H. C. J. Van Beukering and H. Fokker, "Present State-of-the-Art
of the Philips Stirling Engine, " SAE Paper 730646 presented at the
Combined Commercial Vehicle Engineering and Operations, Chicago
(June 1973).
5-4 K. B. United Stirling (Sweden), "Development of Stirling Engines in
Sweden - Review of Recent Progress and Current Activities, " pre-
sented at Conference on Low Pollution, Ann Arbor, Michigan (October
1973).
5-5 "Ford is Readying Stirling Engine for Torino-II, " Automotive Engi-
neering (August 1973).
5-6 Lia Torbjorn, "The Stirling Engine," Combustion Engine Progress
(1973).
5-7 "Low Emission Burner for Rankine Cycle Engines for Automobiles, "
by Solar, Report No. OAWP Publication No. APTD-0707.
5-8 G. T. M. Neelen, et al. , "Stirling Engines in Traction Applications, "
a paper by Philips, United Stirling, and M. A. N. /MWM (no date given;
circa 1970-1972).
R-5
-------
5-9 "Compact Condensers for Rankine Cycle Engines, " by Garrett Cor-
poration, presented at the Automotive Rankine Cycle Contractors
Coordination Meeting (September 1971).
5-10 "C.V.S. Test Simulation of a 128-KW Stirling Passenger Car Engine,"
a report by Philips Research Laboratories (No date given; circa
1970-1972).
5-11 C. W. Rathenau and R. J. Meijer, "Presentation of the Stirling
Engine to Subcommittee on the Environment and the Subcommittee
on Science and Technology and Commerce, United States Senate"
May 1973).
5-12 R. J. Meijer, "Prospects of the Stirling Engine for Vehicular Pro-
pulsion, " Philips Technical Review, Vol. 31, No. 5/6(1970).
5-13 Brochure Material From Philips Laboratories, No. Amer. Philips
Corp. , Briarcliff Manor; New York (November 1973).
6-1 E. F. Obert, Internal Combustion Engines, Third Edition, Interna-
tional Textbook Company, Scranton, Pennsylvania (1968).
6-2 R. C. Bascom, et al. , "Design Factors that Affect Diesel Emissions,"
SAE Paper No. 710484 (June 1971).
6-3 V. S. Yumlu and A. W. Carey, "Exhaust Emission Characteristics
of Four-stroke, Direct-injection, Compression Ignition Engines,"
SAE Paper No. 680420 (May 1968).
6-4 "Cummins Power-Logging, Construction, and Mining", Cummins
Engine Company Bulletin 952790 (February 1972).
6-5 R. Mueller and L. Lacey, "New International Harvester Heavy Duty
Diesel Engines," SAE Paper No. 993A (January 1965).
6-6 R. Malcolm, "International's New Motor Truck V-8 Engines, " SAE
Paper No. 660076 (January 1966).
6-7 H. Ricardo, "The High-speed Internal Combustion Engine, " Blackie
and Son Ltd. , London (1953).
6-8 E. Eisele, "Daimler-Benz Passenger Car Diesel Engines - Highlights
of Thirty Years of Development, " SAE Paper No. 680089 (January
1968).
6-9 Chilton's Automotive Industries; 53rd Annual Engineering Specifica-
tions and Statistical Issue (March 1971).
t>-!0 C. Seippel, "Pressure Exchanger, " U. S. Patent No. 2,399,394(1946).
R-6
-------
6-11 R. C. Weatherston and A. Hertzberg, "The Energy Exchanger, a
New Concept for High Efficiency Gas Turbine Cycles, " Transactions
of the ASME, Journal of Engineering for Power (April 1967).
6-12 Statement by Mercedes Benz of North America; Hearings Before the
Subcommittee on Air and Water Pollution of the Committee on Public
Works, United States Senate (May 18, 1973).
6-13 "Hybrid Heat Engine/Electric Systems Study, " The Aerospace Cor-
poration Report No. TOR-0059(6769-01)-2 (June 1, 1971).
6-14 J. C. Basiletti and E. F. Blackburne, "Recent Developments in
Variable Compression Ratio Engines, " SAE Paper No. 660344 (June
1966).
6-15 Personal Communication with Mr. C. F. Bachle, Consultant to the
EPA, Ann Arbor, Michigan.
6-16 A. Andree and S. J. Pachernegg, "Ignition Conditions in Diesel
Engines," SAE Paper No. 690253 (January 1969).
6-17 W. M. Scott, "Looking in on Diesel Combustion, " SAE Paper No.
690002 (January 1969).
6-18 W. F. Marshall and R. W. Hurn, "Factors Influencing Diesel Emis-
sions," SAE Paper No. 680528 (August 1968).
6-19 S. M. Shahed, et al. , "A Preliminary Model for the Formation of
Nitric Oxide in Direct Injection Diesel Engines and the Application
in Parametric Studies, " SAE Paper No. 730083 (January 1973).
6-20 C. C. J. French, "Taking the Heat Off the Highly Boosted Diesel, "
SAE Paper No. 690463 (August 1969).
6-21 G. J. Barnes, "Relation of Lean Combustion Limits in Diesel Engines
to Exhaust Odor Intensity, " SAE Paper No. 680445 (May 1968).
6-22 B. Lewis and G. Von Elbe, Combustion, Flames and Explosion of
Gases, Academic Press, Inc. , New York (1961^
6-23 C. J. Walder, "Reduction of Emissions from Diesel Engines, " SAE
Paper No. 730214 (January 1973).
6-24 R. E. Bosecker and D. F. Webster, "Precombustion Chamber Diesel
Engine Emissions - A Progress Report, " SAE Paper No. 710672
(August 1971).
R-7
-------
6-25 W- F. Marshall and R. D. Fleming, "Diesel Emissions Re-inventoried, "
Bureau of Mines, Report PB-201896 (July 1971).
6-26 H. A. Ashby, "Final Report, Exhaust Emissions from a Mercedes
Benz Diesel Sedan, " Test and Evaluation Branch, EPA, Ann Arbor,
Michigan (July 1972).
6-27 "Emissions from a Pickup Truck Retrofitted with a Nissan Diesel
Engine, " Emission Control Technology Division, EPA, Ann Arbor,
Michigan (July 1973).
6-28 "Exhaust Emissions from Three Diesel Powered Passenger Cars,"
Emission Control Technology Division, EPA, Ann Arbor, Michigan
(March 1973).
6-29 K. J. Springer, "Emissions from a Gasoline and Diesel Powered
Mercedes 220 Passenger Car, " Southwest Research Institute, San
Antonio, Texas, Report No. AR-813 (June 1971).
6-30 Statement by Peugeot, Inc. ; Hearings Before the Subcommittee on Air
and Water Pollution of the Committee on Public Works, United States
Senate (May 18, 1973).
6-31 Opel Rekord 21 ODD Service Manual (September 1972).
6-32 D. Downs, "The Diesel Engine as a Low Emission Power Unit for
Automobiles, " First Symposium on Low Pollution Power Systems
Development, Ann Arbor, Michigan (October 14-18, 1973).
6-33 K. J. Springer and H. A. Ashby, "The Low Emission Car for 1975 -
Enter the Diesel, " Intersociety Conference on Energy Conversion
(August 1973).
6-34 K. C. Tessier and H. E. Bachman, "Fuel Additives for the Suppres-
sion of Diesel Exhaust Odor and Smoke, " ASME Paper No. 68-WA/
DGP-4 (December 1968).
6-35 K. J. Springer and C. T. Hare, "Four Years of Diesel Odor and
Smoke Control Technology Evaluations - A Summary, " ASME Paper
No. 69-WA/APC-3 (November 1969).
6-36 T. Saito and M. Nabetani, "Surveying Tests of Diesel Smoke Suppres-
sion with Fuel Additives, "SAE Paper No. 730170 (January 1973).
6-37 W. F. Marshall and R. D. Fleming, "Diesel Emissions as Related to
Engine Variables and Fuel Characteristics, " SAE Paper No. 710836
(October 1971).
R-8
-------
6-38 I. M. Khan, et al. , "Factors Affecting Smoke and Gaseous Emissions
from Direct Injection Engines and a Method of Calculation, " SAE
Paper No. 730169 (January 1973).
6-39 Statement by Daimler-Benz Before the Panel oh Environmental
Science and Technology, Subcommittee on Air and Water Pollution,
U. S. Senate Committee on Public Works (March 14, 1972).
6-40 Federal Register, Volume 38, Number 84, Part III, Environmental
Protection Agency (May 2, 1973).
6-41 M. F. Russell, "Reduction of Noise Emissions from Diesel Engine
Surfaces," SAE Paper No. 720135 (January 1972).
6-42 D. D. Tiede and D. F. Kabele, "Diesel Engine Noise Reductions by
Combustion and Structural Modifications, " SAE Paper No. 730245
(January 1973).
6-43 P. S. Myers, et al. , "The ABCs of Engine Exhaust Emissions, " SAE
Paper No. 710481 (June 1971).
6-44 K. J. Springer and R. C. Stahman, "Control of Diesel Exhaust Odors, "
Conference on Odors, The New York Academy of Sciences, New York,
Paper No. 26 (October 1-3, 1973).
6-45 G. J. Barnes, "Relations of Lean Combustion Limits in Diesel Engines
to Exhaust Odor Intensity, " SAE Paper No. 680445 (May 1968).
6-46 "Chemical Analysis of Odor Components in Diesel Exhaust, " A. D.
Little, Inc., Report No. ADL 74744-5 (September 1973).
6-47 K. J. Springer, "An Investigation of Diesel Powered Vehicle Odor and
Smoke - Part III," Southwest Research Report No. AR-695 (October
1969).
7-1 "The Wankel Engine, " Jan Norbye, Chilton Book Company (1971).
7-2 W- G. Froede, "NSU-Wankel Rotating Combustion Engine, " SAE
Paper 288A (January 9-13, 1961).
7-3 "History of Research and Development on Mazda Rotary Engine, " a
brochure published by Toyo Kogyo Limited (June 1970).
7-4 "Mazda Rotary Engine, " a brochure distributed by Toyo Kogyo Limited
(no date).
R-9
-------
7-5 D. E. Cole and C. Jones, "Reduction of Emissions from the Curtiss-
Wright Rotating Combustion Engine with an Exhaust Reactor, " SAE
Paper 700074 (1970).
7-6 "Report By The Committee On Motor Vehicle Emissions," National
Academy of Sciences (February 15, 1973).
7-7 "Alternate Power Sources - Rotary Engine, " Appendix 28 to General
Motors Corporation Request for Suspension of 1976 Federal Emission
Standards, Submitted to EPA (June 1973).
7-8 "Rotary Engine Emissions Development Status, " Ford Motor Company,
Section h, of "Ford 1975-1977 Emission Control Program, Status
Report, " submitted to EPA (November 26, 1973).
7-9 "Rotary Engine Program, " Section 2A7, "Daimler-Benz Request for
Suspension of 1976 Federal Emission Standards," Vol. II, submitted
to EPA (November 1973).
7-10 J. J. Brogan, "Alternative Power Plants, " Division of Advanced Auto-
motive Power Systems, Environmental Protection Agency (15 May 1973).
7-11 Wards Wankel Report, Vol. 2, No. 15 (27 July 1973).
7-12 C. Jones and H. Lamping, "Curtiss-Wright1 s Development Status of
the Stratfield Charge Rotating Combustion Engine, " SAE Paper 710582
(June 1971).
8-1 A. C. Haman, G. E. Cheklick, and A. W- Kaupert, "Review of the
USATACOM Hybrid Combustion Engine Program, " Propulsion Sys-
tems Laboratory, U. S. Army Tank Automotive Command (August
1970).
8-2 J. Baudry, "A New Process for Engine Combustion: A Variable Air-
Fuel Ratio Spark Ignition Engine, " SAE Paper 380 F (1961).
8-3 J. E. Witzky, "Stratified Charge Engines, " ASME Paper 63-MD-42
(May 1963).
8-4 M. Miyake, "Developing a New Stratified Charge Combustion System
with Fuel Injection for Reducing Exhaust Emissions in Small Farm
and Industrial Engines, " SAE Paper 720196 (January 1972).
8-5 J. E. Witzky and J. M. Clark, "A Study of the Swirl Stratified Com-
bustion Principle," SAE Paper 660092 (January 1966).
R-10
-------
8-6 E. Mitchell, J. M. Cobb, and R. A. Frost, "Design and Evaluation
of a Stratified Charge Multifuel Military Engine, " SAE Paper 680042
(January 1968).
8-7 E. Mitchell, et al. , "A Stratified Charge Multifuel Military Engine
A Progress Report," SAE Paper 720051 (January 1972).
8-8 W. J. Coppoc, et al. , "Texaco Controlled Combustion System Pro-
vides an Engine with Clean Exhaust and Good Fuel Economy, " Swedish
Engineering Society, Stockholm, Sweden (March 27, 1973).
8-9 I. N. Bishop and A. Simko, "A New Concept of Stratified Charge Com-
bustion - The Ford Combustion Process (FCP)," SAE Paper 680041
(January 1968).
8-10 A. Simko, M. A. Choma, and L. L. Repko, "Exhaust Emission Con-
trol by the Ford Programmed Combustion Process - PROCO, " SAE
Paper 720052 (January 1972).
8-11 Ford Motor Company, "Request for Suspension of 1976 Motor Vehicle
Exhaust Emission Standards," Volume 1, Sections 1-4 (June 18, 1973).
8-12 J. J. Brogan and G. M. Thur, "Advanced Automotive Power System
Development Program, " 7th Intersociety Energy Conversion Engineer-
ing Conference, San Diego, California (September 25-29, 1972).
8-13 M. Alperstein, Informal Submittal of Test Data by Texaco,
(August 2, 1973).
8-14 H. L. Gompf, "Evaluation of the Texaco Stratified Charge (TCP)
M-151 Army Vehicle, "Environmental Protection Agency, Ann Arbor,
Michigan, Report 73-3 (August 1972).
8-15 "Evaluation of the Texaco Stratified Charge (TCCS) M-151 Army
Vehicle, " Emission Control Technology Division, Environmental Pro-
tection Agency, Ann Arbor, Michigan, Report 73-27 DWP (June 1973).
8-16 Chrysler Corporation, "Application for Syspension of 1976 Motor
Vehicle Emission Standards, Part I" (1973).
8-17 Environmental News, Environmental Protection Agency, Washington,
D. C. (September 24, 1971).
R-ll
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8-18 Federal Register, Part III. Vol. 38, No. 84, Environmental
Protection Agency, Washington, B.C. (May 2, 1973).
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Trends and Influencing Factors, " SAE Paper 730790 (September
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Engines," Contract DAAE 07-70-C-4374, USATACOM (March 1973).
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(granted January 16, 1973).
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93-H9, Washington, D. C. (1973).
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A review for Emission and Efficiency, " Environmental Protection
Agency, for presentation at the 66th Annual Meeting of the APCA,
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Environmental Protection Agency (March 1972).
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2, 884, 913 (granted May 5, 1959).
8-27 "Motor Vehicles, Air Pollution and Health, " A Report of the Surgeon
General to the U. S. Congress, U. S. Department of HEW, Washington,
D. C. (June 1962).
8-28 D. B. Wimmer and R. C. Lee, "An Evaluation of the Performance
and Emissions of a CFR Engine Equipped with a Prechamber, " SAE
Paper 730474 (1973).
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Systems Development, Ann Arbor, Michigan (October 14-19, 1973).
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Designed for Minimum Engine Exhaust Emissions, " SAE Paper
700491 (1970).
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8-31 T. C. Austin, "An Evaluation of Three Honda Compound Vortex
Controlled Combustion (CVCC) Powered Vehicles," Environmental
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8-32 "An Evaluation of a 350 CID Compound Vortex Controlled Combustion
(CVCC) Powered Chevrolet Impala, " Environmental Protection Agency,
Report 74-13 DWP (October 1973).
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western States Section, Ann Arbor, Michigan (Spring 1971).
8-34 R. Becker and R. Gunther, "The Transfer Function of Premixed
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(1968).
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TECHNICAL REPORT DATA
(Please read Instructions on the reverse before completing)
IT NO.
EPA -
. 74. pi 3_h
2.
4. TITLE AND SUBTITLE
Current Status of Alternative Automotive Power
Systems and Fuels
Volume II - Alternative Automotive Engines
3. RECIPIENT'S ACCESSION-NO.
5. REPORT DATE
July 1974
6. PERFORMING ORGANIZATION CODE
D. E. Lapedes, M. G. Hinton, J. Meltzer,
T. lura, O. Dykema, L. Forrest, F. Ghahremani,
W. LPP. W. KnpgglpT- W. Smalley
8. PERFORMING ORGANIZATION REPORT NO.
ATR-74(7325)-l, Vol. II
9. PERFORMING ORGANIZATION NAME AND ADDRESS
The Environmental Programs Group
Environment and Urban Division
The Aerospace Corporation
El Segundo, California 90245
10. PROGRAM ELEMENT NO.
11. CONTRACT/GRANT NO.
68-01-0417
12. SPONSORING AGENCY NAME AND ADDRESS
EPA, Office of Air and Waste Management
Office of Mobile Source Air Pollution Control
Alternative Automotive Power Systems Division
Ann Arbor, Michigan 48105
13. TYPE OF REPORT AND PERIOD COVERED
Final
14. SPONSORING AGENCY CODE
15. SUPPLEMENTARY NOTES
16. ABSTRACT
A summarization has been made of the available nonproprietary information on
the technological status of automotive power systems which are alternatives to the
conventional internal combustion engine, and the technological status of non-
petroleum-based fuels derived from domestic sources which may have application
to future automotive vehicles. The material presented is based principally upon
the results of research and technology activities sponsored under the Alternative
Automotive Power Systems Program which was originated in 1970. Supplementary
data are included from programs sponsored by other government agencies and by
private industry. The results of the study are presented in four volumes; this
volume presents available information pertaining to advanced alternative
automotive heat engines.
KEY WORDS AND DOCUMENT ANALYSIS
DESCRIPTORS
b.IDENTIFIERS/OPEN ENDED TERMS
c. COSATl l;icld/Group
Automotive Heat Engines
Gas Turbine
Rankine Cycle
Stirling Cycle
Diesel
Wankel (Rotary Piston)
Stratified Charge
Exhaust Emissions
Fuel Economy
Technology Status
8. DISTRIBUTION STATEMENT
Unlimited
19. SECURITY CLASS (This Report;
Unclassified
O. OF PA
463
20. SECURITY CLASS (This page/
Unclassified
22. PRICE
EPA Form 2220-1 (9-73)
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