PERFORMANCE OF A DUAL-FUEL PRECHAMBER

DIESEL E N G I N E WITH NATURAL GAS

BY

SEAHO (SONG

.Sc., The U n i v e r s i t y of B r i t i s h Columbia, 1

A T H E S I S SUBMITTED IN P A R T I A L FULFILLMENT

THE REQUIREMENTS FOR THE DEGREE OF

MASTER OF APPLIED SCIENCE

in

THE F A C U L T Y OF GRADUATE STUDIES

THE DEPARTMENT OF M E C H A N I C A L ENGINEERING

We accept this thesis as conforming

to the required standard

THE UNIVERSITY OF B R I T I S H COLUMBIA

May 1984

©Seaho Song, 1984

In p r e s e n t i n g t h i s thesis in p a r t i a l fulfilment of the
r e q u i r e m e n t s f o r an advanced degree at the U n i v e r s i t y of
B r i t i s h Columbia, I agree t h a t the L i b r a r y s h a l l make i t
freely available for reference and study. I further
agree t h a t permisiion for extensive copying of this
t h e s i s f o r s c h o l a r l y purposes may be g r a n t e d by the Head
of my Department or by h i s or her representatives. It
i s u n d e r s t o o d t h a t c o p y i n g or p u b l i c a t i o n of t h i s thesis
for financial gain shall not be allowed without my
written permission.

Department of M e c h a n i c a l E n g i n e e r i n g

The U n i v e r s i t y of B r i t i s h Columbia
2324 Main M a l l
Vancouver, Canada
V6T 1W5

Date: June 1984

 ii

ABSTRACT

The f e a s i b i l i t y of d u a l - f u e l o p e r a t i o n w i t h n a t u r a l gas i n
a prechamber d i e s e l engine was s t u d i e d w i t h s p e c i a l emphasis on
f u e l consumption and c y l i n d e r p r e s s u r e development. The e f f e c t s
of a i r r e s t r i c t i o n , p i l o t d i e s e l flow r a t e and i n j e c t i o n timing
were also studied. D u a l - f u e l o p e r a t i o n showed poor p a r t - l o a d
f u e l consumption; near f u l l l o a d t h e f u e l consumption was close
to that of straight diesel operation. In t h e absence of
i n j e c t i o n t i m i n g adjustment t h e maximum power output of dual-
fuel operation was severely limited by t h e maximum c y l i n d e r
pressure. Retarding the injection timing was effective in
reducing t h e maximum cylinder pressure to a safe l e v e l . The
a n a l y s i s of apparent energy r e l e a s e i n d i c a t e s t h e d i f f e r e n c e s i n
combustion mechanism between a u t o - i g n i t i o n of diesel fuel in
straight diesel operation and p r o p a g a t i o n of flame f r o n t s i n
dual-fuel operation.
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T a b l e of Contents

Abstract i i
L i s t of T a b l e s iv
L i s t of F i g u r e s y
Acknowledgements viii
Nomenclature ix
I . INTRODUCTION 1
1 .1 Background - 1
1.2 P r e s e n t Study 7
I I . REVIEW OF LITERATURE 10
2.1 H i s t o r y of D u a l - F u e l D i e s e l Engine 10
2.2 Review of Research 12
I I I . APPARATUS AND INSTRUMENTATION ... 26
3.1 Engine and Test Bed • 26
3.2 I n s t r u m e n t a t i o n 38
3.3 F u e l 43
3.4 Data P r o c e s s 45
IV. EXPERIMENTAL RESULTS 47
4.1 F u e l Consumption 47
4.1.1 F u e l Consumption w i t h U n m o d i f i e d Engine .. 47
4.1.2 E f f e c t of R e s t r i c t i n g I n t a k e A i r 55
4.1.3 E f f e c t of V a r y i n g I n j e c t i o n Timing 56
4.2 C y l i n d e r P r e s s u r e 63
4.2.1 C y l i n d e r P r e s s u r e i n U n m o d i f i e d Engine ... 63
4.2.2 E f f e c t of R e s t r i c t i n g I n t a k e A i r 74
• 4.2.3 E f f e c t of V a r y i n g I n j e c t i o n T i m i n g 79
V. ANALYSIS OF APPARENT ENERGY RELEASE 84
5.1 G e n e r a l 84
5.2 Method of C a l c u l a t i o n 85
5.2.1 D e f i n i t i o n s , E q u a t i o n s , and Assumptions .. 85
5.2.2 Computation Procedure 98
5.3 A n a l y s i s 107
5.3.1 O p e r a t i o n w i t h Unmodified Engine 107
5.3.2 E f f e c t of R e s t r i c t i n g I n t a k e A i r 124
5.3.3 E f f e c t of V a r y i n g I n j e c t i o n T i m i n g ....... 130
V I . CONCLUSIONS AND RECOMMENDATIONS 135
6.1 C o n c l u s i o n s 135
6.2 Recommendations 138
BIBLIOGRAPHY 139
APPENDIX A - CALIBRATION CURVES 142
APPENDIX B - COMPUTATION OF INDICATED MEAN EFFECTIVE
PRESSURE 146
APPENDIX C - COMPUTER PROGRAM FOR DATA ACQUISITION .... 148
APPENDIX D - COMPUTER PROGRAM FOR DATA PROCESS 157
APPENDIX E - COMPUTER PROGRAM FOR APPARENT ENERGY
RELEASE 162
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L i s t of Tables

2.1 Summary of Past E x p e r i m e n t a l Work 13

3.1 Engine S p e c i f i c a t i o n 29
3.2 T y p i c a l c o m p o s i t i o n of the N a t u r a l Gas Used 44
3.3 T y p i c a l Output of Computer Program f o r Data
Processing 46
5.1 Comparison of A c t u a l and Computed F u e l Energy
Consumed . . 105
5.2 E f f e c t of I n t a k e A i r R e s t r i c t i o n on M i x t u r e
Temperature a t Top Dead Center 126
 V

L i s t of F i g u r e s

1.1 Combustion Chambers o f . D i r e c t - I n j e c t i o n and

Prechamber Engines 5
2.1 E f f e c t of G a s - A i r M i x t u r e S t r e n g t h on I g n i t i o n
delay 17
2.2 T y p i c a l Pressure-Time Trace of Non-Knocking and
Knocking O p e r a t i o n . . w • 21
2.3 V a r i a t i o n of Power Output w i t h t h e O v e r a l l M i x t u r e
S t r e n g t h f o r D i f f e r e n t I n t a k e Tempertures 22
2.4 T y p i c a l Thermal E f f i c i e n c i e s of D u a l - F u e l and
Straight D i e s e l Operation 25
3.1 Apparatus and I n s t r u m e n t a t i o n 27
3.2 Flow of A i r , F u e l , and Exhaust Gas 28
3.3 Shape of Combustion Chambers 30
3.4 Sleeve M e t e r i n g F u e l System 32
3.5 F u e l I n j e c t i o n Pump and Housing 32
3.6 Seqence of I n j e c t i o n E v e n t s 33
3.7 Governor Components of S l e e v e M e t e r i n g 33
3.8 Fuel I n j e c t i o n Nozzle 34
3.9 Turbocharger Cutaway View 36
3.10 Gas M i x e r 37
3.11 Mounting of C y l i n d e r P r e s s u r e Transducer 39
4.1 E f f e c t of P i l o t D i e s e l Flow Rate on Brake Thermal
Efficiency 48
4.2 F u e l Consumption a t i d l i n g O p e r a t i o n 50
4.3 E f f e c t of p i l o t D i e s e l Flow Rate on I n d i c a t e d
Thermal E f f i c i e n c y 52
4.4 Comparison of Brake Thermal e f f i c i e n c i e s f o r
D u a l - F u e l and S t r a i g h t D i e s e l O p e r a t i o n 54
4.5 E f f e c t of I n t a k e A i r R e s t r i c t i o n on Brake
Thermal E f f i c i e n c y 57
4.6 T y p i c a l C y l i n d e r P r e s s u r e Trace and Apparent
P o i n t of I g n i t i o n S t a r t 58
4.7 Apparent P o i n t of I g n i t i o n S t a r t a t V a r i o u s
Loads 59
4.8 E f f e c t of V a r y i n g I n j e c t i o n Timing on Brake
Thermal E f f i c i e n c y 61
4.9 P-V Diagram of S t r a i g h t D i e s e l O p e r a t i o n 64
4.10 Ln P-V Diagram of S t r a i g h t D i e s e l O p e r a t i o n 66
4.11 Comparison of P-V Diagrams f o r D u a l - F u e l
and S t r a i g h t D i e s e l O p e r a t i o n 67
4.12 Comparison of Ln P-V Diagrams f o r D u a l - F u e l
and S t r a i g h t D i e s e l O p e r a t i o n 69
4.13 Comparison of Maximum C y l i n d e r P r e s s u r e s f o r
D u a l - F u e l and S t r a i g h t D i e s e l O p e r a t i o n 70
4.14 Maximum C y l i n d e r P r e s s u r e a t V a r i o u s Loads 71
 vi

4.15 Comparison of Maximum Rate of C y l i n d e r P r e s s u r e

R i s e f o r D u a l - F u e l and S t r a i g h t D i e s e l
Operation 72
4.16 Maximum Rate of C y l i n d e r P r e s s u r e R i s e a t
V a r i o u s Loads 73
4.17 E f f e c t of I n t a k e A i r R e s t r i c t i o n on Maximum
Cylinder Pressure 75
4.18 E f f e c t of I n t a k e A i r R e s t r i c t i o n P r e s s u r e P r i o r
to Combustion 76
4.19 E f f e c t of I n t a k e A i r R e s t r i c t i o n on Maximum
Rate of C y l i n d e r P r e s s u r e R i s e 78
4.20 E f f e c t of V a r y i n g I n j e c t i o n Timing on Maximum
c y l i n d e r P r e s s u r e and Rate of P r e s s u r e R i s e 80
4.21 E f f e c t of V a r y i n g I n j e c t i o n Timing on P r e s s u r e
P r i o r t o Combustion and P o i n t of I g n i t i o n S t a r t .. 81
4.22 E f f e c t of V a r y i n g I n j e c t i o n Timing on P-V
Diagram 82
5.1 C o n t r o l Volume f o r Apparent Energy R e l e a s e
Analysis 86
5.2 Apparent Heat T r a n s f e r Rate and Heat T r a n s f e r
Model 91
5.3 E f f e c t of Heat T r a n s f e r Model on Apparent Rate
of Energy Release 93
5.4 E f f e c t of E q u i l i b r i u m C a l c u l a t i o n on Apparent
Rate of Energy Release 97
5.5 E f f e c t of Smoothing P r e s s u r e Data on Apparent
r a t e of Energy Release 101
5.6 F l o w c h a r t of Computer Program f o r Apparent Energy
Release 103
5.7 T y p i c a l Output of Computer Program f o r Apparent
Energy R e l e a s e A n a l y s i s 104
5.8 Rate of Energy Release of S t r a i g h t D i e s e l
O p e r a t i o n a t V a r i o u s Loads 108
5.9 Cumulative Energy Release of S t r a i g h t D i e s e l
O p e r a t i o n a t V a r i o u s Loads 109
5.10 E f f e c t of A i r - F u e l R a t i o on Maximum Rate of
Energy R e l e a s e i n S t r a i g h t D i e s e l O p e r a t i o n 111
5.11 Rate of Energy Release of D u a l - F u e l O p e r a t i o n a t
V a r i o u s Loads 112
5.12 E f f e c t of G a s - A i r M i x t u r e S t r e n g t h on Maximum
Rate of Energy R e l e a s e i n D u a l - F u e l O p e r a t i o n .... 113
5.13 Cumulative Energy R e l e s e of D u a l - F u e l O p e r a t i o n
at V a r i o u s Loads 115
5.14 Comparison of Rate of Energy R e l e a s e f o r
S t r a i g h t D i e s e l and D u a l - F u e l O p e r a t i o n 116
5.15 Comparison of Cumulative Energy R e l e a s e f o r
S t r a i g h t D i e s e l and D u a l - F u e l O p e r a t i o n 117
5.16 Rate of Energy R e l e a s e of D u a l - F u e l O p e r a t i o n
at V a r i o u s P i l o t D i e s e l Flow Rates 118
 vii

5.17 Cumulative Energy Release of D u a l - F u e l O p e r a t i o n

at V a r i o u s P i l o t D i e s e l Flow Rats 120
5.18 F r a c t i o n of F u e l Burnt i n Low Load D u a l - F u e l
Operation 121
5.19 Rate of Energy R e l e a s e of D u a l - F u e l O p e r a t i o n
at V a r i o u s P i l o t D i e s e l Flow Rates 122
5.20 C u m u l a t i v e Energy Release of D u a l - F u e l O p e r a t i o n
at V a r i o u s P i l o t D i e s e l Flow Rates 123
5.21 E f f e c t of R e s t r i c t i n g I n t a k e A i r on Rate of
Energy R e l e a s e 125
5.22 E f f e c t of R e s t r i c t i n g I n t a k e A i r on C u m u l a t i v e
Energy Release 127
5.23 E f f e c t of R e s t r i c t i n g I n t a k e A i r on Rate of
Energy Release 128
5.24 E f f e c t of R e s t r i c t i n g I n t a k e A i r on C u m u l a t i v e
Energy Release 129
5.25 E f f e c t of Advancing I n j e c t i o n Timing on Rate of
Energy Release 131
5.26 E f f e c t of Advancing I n j e c t i o n Timing on
Cumulative Energy Release 132
5.27 E f f e c t of R e t a r d i n g I n j e c t i o n Timing on Rate of
Energy Release 133
5.28 E f f e c t of R e t a r d i n g I n j e c t i o n Timing on
Cumulative Energy Release 134

\
 vi i i

Acknowledgement

The author wishes t o acknowledge a sincere gratitude to

Dr. P.G. H i l l for his advice and encouragement. Numerous
d i s c u s s i o n s w i t h Dr. Roger M i l a n e have and w i l l remain valuable
to the author. Thanks are a l s o due t o Messrs Stan Mah, Shu
Osaka, and John Hoar f o r technical advice and assistance in
setting up the equipments. Stan Mah was r e s p o n s i b l e f o r the
installation of the engine, and for development of the
instrumentation before the project commenced. P r o v i s i o n of
t e c h n i c a l i n f o r m a t i o n and a s s i s t a n c e by Mr. J i m Bare of F i n n i n g s
Tractors i s greatly appreciated. F u r t h e r thanks a r e due t o the
members of the t h e s i s committee, Dr. B. A h l b o r n , Dr. B. Evans
and Dr. E.G. Hauptmann.

This work was financially supported by the Federal

Department of Energy Mines and R e s o u r c e s .
 ix

Nomenclature

A area m 2

ATDC after top dead center

BMEP brake mean e f f e c t i v e pressure kPa
BTDC before top dead center
CA crank angle
CE chemical energy kJ
D bore mm
E energy kJ
e c internal energy of combustion kJ/kg
K equilibrium constant
k thermal conductivity kW/(irt' C)
m^ mass of fuel kg
N number of moles kmole
P pressure MPa
Q heat transfer kJ
q heat transfer rate kW
R Reynolds Number
T temperture C
U internal energy kJ
V volume m 3

V mean velocity m/s

W work kJ
V equilibrium composition
v viscosity kg/m-s
p density kg/m 3

$ equivalence r a t i o
X inverse of equivalence r a t i o

Subscripts:
9 gas
i state
 1

CH.I Introduction

1 .1 Background

Dual-Fuel D i e s e l Engines w i t h N a t u r a l Gas

D u a l - f u e l d i e s e l engines a r e here d e f i n e d as those which

burn e i t h e r gaseous f u e l s or d i e s e l , or both a t t h e same time.
The mode of o p e r a t i o n i s defined as straight diesel i f only
diesel f u e l i s used, and d u a l - f u e l i f two f u e l s a r e used a t the
same t i m e . In d u a l - f u e l o p e r a t i o n the gaseous fuel i s mixed
with a i r at lean gas-air ratio and the m i x t u r e i s then
compressed d u r i n g t h e compression s t r o k e . Near the end of the
compression stroke, diesel f u e l i s i n j e c t e d and a u t o - i g n i t e s ,
i n i t i a t i n g the combustion of the g a s - a i r mixture. Because of
its f u n c t i o n t o i n i t i a t e the combustion the d i e s e l i n d u a l - f u e l
o p e r a t i o n i s o f t e n r e f e r r e d t o as p i l o t d i e s e l . The changeover
of t h e mode of operation, e i t h e r from d u a l - f u e l t o s t r a i g h t
d i e s e l o r s t r a i g h t d i e s e l t o d u a l - f u e l can take p l a c e w h i l e the
engine operates.

Combustion characteristics of dual-fuel operation differ

from t h o s e of s t r a i g h t d i e s e l o p e r a t i o n . In d i e s e l operation

the combustion t a k e s p l a c e s w i t h i n s m a l l zones where the f u e l -

a i r r a t i o i s s u i t a b l e f o r combustion. As a stream of diesel

fuel i s i n j e c t e d i n t o t h e c y l i n d e r , i t i s mixed w i t h a i r t o be

d i s i n t e g r a t e d i n t o f i n e d r o p l e t s which i n turn vapourize and

 2

auto-ignite due t o the h i g h temperature of the compressed a i r .

The time p e r i o d d u r i n g which l i q u i d d i e s e l i s mixed w i t h a i r and
v a p o u r i z e d i s r e f e r r e d t o as ' p h y s i c a l d e l a y ' and the time taken
from then t o the p o i n t j u s t p r i o r t o i g n i t i o n i s r e f e r r e d t o as
'chemical delay'. These two delay p e r i o d s a r e combined and
commonly termed 'ignition delay'. Combustion in dual-fuel
operation, i n c o n t r a s t , o c c u r s i n a n e a r l y homogeneous f u e l - a i r
mixture. D u r i n g the i n t a k e s t r o k e , a n e a r l y u n i f o r m m i x t u r e of
gas and air is drawn into the c y l i n d e r , then compressed by
p i s t o n movement t o h i g h temperature and p r e s s u r e but not high
enough to e l i c i t auto-ignition. A s m a l l amount of d i e s e l fuel
i s i n j e c t e d i n t o the homogeneous g a s - a i r m i x t u r e near the end of
the compression s t r o k e . The injected pilot diesel subsequently
goes through the ignition delay before i t d i s i n t e g r a t e s into
d i e s e l ' vapour t o i n i t i a t e flame f r o n t s which propagate through
the g a s - a i r m i x t u r e . The p r o p a g a t i o n of flame f r o n t s i s l a r g e l y
responsible for subsequent combustion of the r e m a i n i n g g a s - a i r
mixture. I t i s i n t h i s r e g a r d t h a t the combustion process of
dual-fuel operation differs from that of straight diesel
operation. The combustion in straight diesel operation is
largely due to auto-ignition of d i e s e l f u e l , whereas t h a t of
d u a l - f u e l o p e r a t i o n depends h e a v i l y on both the auto-ignition
characteristics of pilot diesel and the p r o p a g a t i o n of flame
fronts.

The gaseous f u e l we a r e concerned w i t h here i s n a t u r a l gas.

N a t u r a l gas i s a v a i l a b l e i n most localities, and is in many

places more abundant than o t h e r k i n d s of f u e l . The temperature

at which the n a t u r a l gas a u t o - i g n i t e s i s higher than that of

 3

other available gaseous fuels. Because of the high auto-

ignition temperature the gas-air mixture can be compressed to

high compression ratio without auto-ignition. This, with the

low cost of natural gas, makes dual-fuel diesel operation with

natural gas an a t t r a c t i v e means for power production, especially

in places where the gas supply may possibly be interrupted.

Dual-fuel diesel engines with natural gas have been

employed extensively in power generating stations where engines

operate constantly at near full load. The fuel consumption rate

of dual-fuel diesel operation with natural gas at full load has

been shown to be as good as or sometimes better than that of

straight diesel operation. Typically, the amount of pilot

diesel used in the past has consisted of less than 10% of total

energy input. In applications such as pipe line industries,

where the required power is nearly constant, dual-fuel diesel

operation has been shown to be satisfactory.

The maximum power of dual-fuel diesel operation is limited

by the occurrence of knock with its high rate of cylinder

pressure rise. Knock in dual-fuel engines is believed to be the

same type as in spark-ignition engine. If the dual-fuel engine

has to operate over a range of load, the economic advantage of

dual-fuel diesel operation at full load may be offset by poor

combustion c h a r a c t e r i s t i c s at low load, which result in poor

fuel consumption rate. At low load significant amounts of

natural gas survive combustion and escape through the exhaust

because of lean gas-air mixture strength. Low load fuel

consumption rate can be improved by restricting intake air,

 4

which effectively increases gas-air mixture s t r e n g t h , or by

increasing the pilot diesel flow r a t e . Other possible methods
which improve low l o a d f u e l consumption r a t e are preheating the
i n t a k e charge and advancing the injection timing of pilot
diesel. P r e h e a t i n g the i n t a k e charge r e s u l t s in higher mixture
temperature and thus assists the oxidation reaction of the
gaseous fuel. Advancing the i n j e c t i o n t i m i n g p r o v i d e s longer
time f o r the gaseous f u e l to r e a c t subsequent to the initial
reaction of the pilot diesel.

Prechamber D i e s e l Engine

Diesel engines can be c l a s s i f i e d i n t o two types: d i r e c t -

i n j e c t i o n or i n d i r e c t - i n j e c t i o n t y p e s depending on whether the
combustion involves one or two chambers. A prechamber diesel
engine is an indirect-injection engine consisting of two
chambers, prechamber and main chamber. Fig. 1.1 shows the
t y p i c a l shape of the combustion chambers for direct-injection
and prechamber d i e s e l e n g i n e s . The volume of the prechamber i s
t y p i c a l l y 20 t o 30 p e r c e n t of the c l e a r a n c e volume. The main
objective of the prechamber d e s i g n i s to burn a s m a l l fraction
of the injected f u e l i n the prechamber so that the resulting
pressure, r i s e will drive the m i x t u r e of p a r t i a l l y burnt and
unburned f u e l i n t o the main chamber as a h i g h speed jet whose
t u r b u l e n c e w i l l promote r a p i d and complete combustion.

The advantages of the prechamber d i e s e l engine compared to

the d i r e c t - i n j e c t i o n type are better emission characteristics

F i g u r e 1 .1 - Combustion Chambers of D i r e c t - I n j e c t i o n and Prechamber Engines
 6

and less tendency to knock. Because of the b e t t e r emmision

c h a r a c t e r i s t i c s the prechamber engines are p r e f e r r e d f o r higher
speed operations. The d i s a d v a n t a g e s a r e m a i n l y a s s o c i a t e d w i t h
higher surface-to-volume r a t i o which enhances heat loss and
t h r o t t l i n g between prechamber and main chamber. These r e s u l t i n
h i g h e r f u e l consumption r a t e .

The type of d i e s e l engine used i n d u a l - f u e l o p e r a t i o n with

n a t u r a l gas has been almost exclusively direct-injection, so
that the behaviour of a prechamber d i e s e l engine w i t h d u a l -
f u e l l i n g over a range of l o a d appears t o be v i r t u a l l y unknown.
 7

1.2 P r e s e n t Study

Object ives

The p r i m a r y o b j e c t i v e of t h i s study was to determine the

feasibility of dual-fuel operation with natural gas in a
prechamber diesel engine. Observations were made of fuel
consumption and cylinder pressure development which may
critically affect engine durability. The effects on fuel
consumption and cylinder pressure of the f o l l o w i n g variables
were s t u d i e d :

a. flow r a t e of p i l o t diesel
b. g a s - a i r m i x t u r e s t r e n g t h
c. i n j e c t i o n t i m i n g of p i l o t diesel
Computer a n a l y s i s of apparent energy release was employed to
study the combustion characteristics, first for straight diesel
o p e r a t i o n and then w i t h r e g a r d to the above three operating
variables. Past experience w i t h d i r e c t - i n j e c t i o n engines was
considered in a n t i c i p a t i n g possible operating difficulties and
i n a n a l y z i n g the observed combustion characteristics.

E x p e r i m e n t a l Work

A C a t e r p i l l a r 3304, f o u r - c y l i n d e r prechamber marine e n g i n e ,

with turbocharger was used in the course of the study. The

e x p e r i m e n t a l study was done at constant engine speed because

load changes r a t h e r than speed changes were c o n s i d e r e d t o be of

 8

chief concern i n examining the feasibility of dual-fuel

operation. Throughout t h e e x p e r i m e n t s measurments were taken t o
produce f u e l consumption and c y l i n d e r p r e s s u r e data.

The first phase of t e s t s was performed w i t h v a r i o u s loads

and p i l o t d i e s e l fuel rates. I t was found that the fuel
consumption rate of d u a l - f u e l o p e r a t i o n i s considerably higher
than t h a t of s t r a i g h t d i e s e l o p e r a t i o n a t loads much lower than
the 'full load' s p e c i f i e d by t h e engine m a n u f a c t u r e r . As the
l o a d was i n c r e a s e d t o near f u l l load the f u e l consumption rate
approached that of s t r a i g h t d i e s e l o p e r a t i o n . I t was observed
t h a t a t some p o i n t beyond f u l l l o a d t h e f u e l consumption r a t e of
the d u a l - f u e l o p e r a t i o n would become lower than t h a t of s t r a i g h t
diesel operation. The flow r a t e of p i l o t d i e s e l was shown t o
greatly affect the f u e l consumption r a t e a t low l o a d s . As t h e
l o a d was i n c r e a s e d the e f f e c t of p i l o t d i e s e l flow rate became
smaller. The maximum c y l i n d e r p r e s s u r e and p r e s s u r e r i s e were
observed t o i n c r e a s e r a p i d l y with increase i n load. .The maximum
power output was s e v e r e l y ' l i m i t e d •by the maximum cylinder
pressure. I t was found t h a t when t h e flow r a t e of p i l o t d i e s e l
was below a c e r t a i n l i m i t the operation became erratic with
misfirings. F o r a range of l o a d s and p i l o t d i e s e l f l o w r a t e s , a
region of unstable operation due t o i n s u f f i c i e n t p i l o t d i e s e l
flow r a t e was e s t a b l i s h e d .

The second phase of t h e e x p e r i m e n t s was i n t e n d e d f o r study

of the effect of intake a i r r e s t r i c t i o n . During the i n i t i a l

stage of t h e e x p e r i m e n t s i t was noticed that excessive a i r

restriction can cause surge of the compressor i n the

 9

turbocharger. S i n c e surge can e a s i l y cause m e c h a n i c a l damage of

the t u r b o c h a r g e r t h e i n t a k e a i r r e s t r i c t i o n had t o be l i m i t e d t o
a s m a l l range of a i r f l o w r e d u c t i o n .

In the l a s t stage of experiments t h e e f f e c t of injection

timing was studied primarily because of concern over maximum
pressures associated with dual-fuelling and normal injection
timing. I t was found that retarding the injection timing
r e s u l t s i n s i g n i f i c a n t r e d u c t i o n i n both t h e maximum cylinder
p r e s s u r e and p r e s s u r e r i s e . The change i n f u e l consumption rate
due t o the r e t a r d e d i n j e c t i o n was found t o be s m a l l .

Computer A n a l y s i s of Apparent Energy Release

A computer program which computes apparent energy release

due t o combustion was developed i n order t o study the combustion
characteristics. The c y l i n d e r p r e s s u r e d a t a o b t a i n e d d u r i n g the
course of above t h r e e phases of e x p e r i m e n t s were used i n the
analysis. The analysis showed that the e x c e s s i v e maximum
cylinder pressure was mainly associated with high rate of
combustion energy release. The results of t h e a n a l y s i s a r e
consistent with different mechanisms of combustion: auto-
ignition of d i e s e l i n s t r a i g h t d i e s e l o p e r a t i o n and p r o p a g a t i o n
of flame f r o n t s i n d u a l - f u e l operation.
 10

CH.II Review of L i t e r a t u r e

2.1 H i s t o r y of t h e D u a l - F u e l D i e s e l Engine

The e a r l i e s t p r a c t i c a l use of gas as an engine fuel dates

back to the end of the 19th c e n t u r y . S p a r k - i g n i t e d engines
called 'gas e n g i n e s ' o p e r a t e d i n much t h e same way as modern gas
engines. In t h e s e early engines the g a s - a i r mixtures were
n e a r l y s t o i c h i o m e t r i c and t h e compression r a t i o s were about 6:1.
Commercial p r o d u c t i o n of e n g i n e s of v a r i o u s s i z e s began i n about
the year 1900. Jones(l944) states that by 1920 i n B r i t a i n
engines w i t h maximum power r a n g i n g from 5 t o 2000 horse power (4
to 1500 kW) were manufactured f o r use mainly on waste gases,
p a r t i c u l a r l y b l a s t furnace gas.

The first attempt to burn gas i n a compression ignition

engine appears t o have been made by the C.&G. Cooper Company in
1927. According to Boyer and Crooks(1951), the f i r s t test
c o n s i s t e d of i n j e c t i n g n a t u r a l gas a l o n e a t h i g h p r e s s u r e a t the
end of t h e compression stroke. This resulted in irregular
firing of t h e gas. In t h e next t e s t a s m a l l p o r t i o n of d i e s e l
was i n j e c t e d i n a d d i t i o n t o t h e gas. T h i s was t h e b i r t h of the
so called 'gas-diesel' engine, i n which t h e gaseous f u e l was
i n j e c t e d i n t o the c y l i n d e r a t h i g h p r e s s u r e (about 1000 p s i or
7 MPa). The first commercial i n s t a l l a t i o n of such engine was
a c h i e v e d by t h e Nordberg Company i n 1935. A 1,665 horse power
(1,241 kW) engine was installed at Lubbock, Texas, and the
operation was successful; at full load the specific fuel
 11

consumption was as low as t h a t of d i e s e l o p e r a t i o n .

The high pressure gas i n j e c t i o n equipment needed f o r the

' g a s - d i e s e l ' engines gave r i s e t o problems: the equipment was
costly and thus limited to use on l a r g e e n g i n e s , and i t was
rather d i f f i c u l t to m a i n t a i n . In 1938 the N a t i o n a l Gas and Oil
Company developed an 8 - c y l i n d e r 440 horse power e n g i n e , which
used town gas. In t h i s engine the gas was admitted to the
c y l i n d e r a t low p r e s s u r e through a s e p a r a t e passage from the a i r
inlet. The o p e r a t i o n was s u c c e s s f u l and l e d to the conversion
of e x i s t i n g engines at the C o l e s h i l l Works of the Birmingham
Tame and Rea D i s t r i c t Drainage Board.

An alternative means of a d m i t t i n g the gas at low pressure

became commonly used. M i t c h e l l and Whitehouse(1954) described
one such scheme developed by the E n g l i s h E l e c t r i c Company t o
provide better gas-air mixing. Instead of being directly
admitted i n t o the c y l i n d e r , gas was mixed w i t h a i r i n the i n t a k e
manifold after upstream i n j e c t i o n through a 'flutter' valve.
T h i s f l u t t e r v a l v e a c t e d as n o n - r e t u r n v a l v e t o prevent pressure
p u l s a t i o n s or e x p l o s i o n s p a s s i n g back i n t o the gas supply pipes.

The literature of the 1950's reveals some fundamental

studies of dual-fuel operation on direct-injection diesel
engines. These are discussed in detail in the following
section.
 12

2.2 Review of Research

An extensive summary of the r e s u l t s of d u a l - f u e l d i e s e l

combustion r e s e a r c h i s p r o v i d e d i n review papers by Karim(l980)
and Karim(l982). Both r e s e a r c h r e s u l t s and a p p l i c a t i o n s have
been reviewed by 0'Neal(1982). The literature contains a
considerable amount of information concerning the operating
experience and combustion processes of dual-fuel diesel
operation though t h i s i s r e s t r i c t e d t o d i r e c t - i n j e c t i o n engines
only. The research experiences are d i s c u s s e d i n t h i s s e c t i o n i n
c h r o n o l o g i c a l order as i n d i c a t e d i n T a b l e 2.1 which mentions the
main f e a t u r e s of each p r o j e c t . The review here i s r e s t r i c t e d t o
s t u d i e s which i n v o l v e methane-based gases.

The importance of g a s - a i r mixture strength in dual-fuel

operation was studied by Elliott & Davis(l95l). In t h e i r
e x p e r i m e n t s w i t h a CFR" d i e s e l engine a t a compression r a t i o of
21:1, selected pilot diesel rates were held constant to
determine the e f f e c t of the c o n c e n t r a t i o n of n a t u r a l gas (88.9%
methane, 10.6% other hydrocarbons) in the intake. Their
experiments showed that when the g a s - a i r m i x t u r e strength was
below a c e r t a i n l i m i t , the p r o p o r t i o n of gas reacting increased
with diesel fuel-air r a t i o and gas-air mixture strength. They
found t h a t i f the c o n c e n t r a t i o n of gas i s below the lower limit
of f l a m m a b i l i t y (which i s commonly d e f i n e d as the concentration

*CFR ( C o o p e r a t i v e F u e l Research) e n g i n e :
single-cylinder engine, with 3.25 i n . (82.5 mm) bore and
4.5 i n . (114 mm) s t r o k e , manufactured by the Waukeska Engine
Co. of Waukeska, Wis., the s t a n d a r d engine used f o r d e t o n a t i o n
measurement and g e n e r a l l y f o r d e t o n a t i o n r e s e a r c h .
 AUTHOR DATE ENGINE GASEOUS FUEL MAIN FINDINGS

Elliot CFR diesel n a t u r a l gas l o w e r l i m i t o f f lammab i 1 i ty,

& 1951 c r . 16: 1, 21: 1 methane 88.9% d e p e n d e n c e o f amount o f g a s
Davis other r e a c t e d on g a s - a i r m i x t u r e
hydrocarbon 10.6% s t r e n g t h and d i e s e l fuel-
air ratio.

single-cylinder s l u d g e gas effect of intake a i r r e -

Lewis 1954 direct-injection me t h a n e 8 6.8% s t r i c t i o n and i n t a k e a i r
c r . 14.7:1 nitrgen 4.5% preheating on e f f i c i e n c y .
b o r e 105mm carbon
s t r o k e 152mm dioxide 5.5%

e f f e c t of intake a i r pre-
S i m o n s on 1954 single-cylinder me t h a n e h e a t i n g on r e a c t i o n s o f
direct-injection gaseous charge.
c r . 14.7:1

single-cylinder e f f e c t of intake a i r pre-

S i m o n s on 1955 direct-injection methane heating, intake a i r r e s t r i -
c r . 14.7:1 c t i o n , and v a r y i n g inject-
b o r e 105mm i o n t i m i n g on e f f i c i e n c y .
s t r o k e 152mm

Moore single-cylinder e f f e c t of intake a i r pre-

& 1955 direct-injection me t h a n e heating, intake a i r restr-
Mitchell bore 105mm i c t i o n , and v a r y i n g inject-
s t r o k e 152mm i o n t i m i n g on e f f i c i e n c y .

Mi t c h e l l f our-cylinder sludge gas e f f e c t of intake a i r r e -

& 1955 direct-injection methane 87.9% s t r i c t i o n on e f f i c i e n c y
Wh i t e h o u s e c r. 13.5:1 nitrogen 4.4%
bore 254mm carbon
s t r o k e 30 5mm dioxide 5.5%
 Felt single-cylinder n a t u r a l gas p r o b l e m s w i t h l o s s o f com-
& 1962 direct-injection me t h a n e 8 7.1% bustion c o n t r o l , e f f e c t of
Steele c r . 16.2:1 nitrogen 7 . 1% additives on k n o c k - l i m i t e d
other power.
hydrocarbon 5.1%

Kar im, single-cylinder determination of knock-

Klat, 1966/67 direct-injection me t h a n e 9 7.8% l i m i t e d powe r .
& c r . 14.2:1
Moore bore 108mm
s t r o k e 152mm

K a r im single-cylinder heat r e l e a s e analysis,

& 1968 direct-injection me t h a n e two-phased combustion.
Kahn bore 105mm
s t r o k e 152mm

* gas c o m p o s i t i o n based on v o l u m e

Table 2.1 - Summary o f Past Experimental Work

 15

of gaseous fuel i n the i n t a k e a i r at which the minimum l i q u i d

fuel-air ratio yields consistent and close-to-complete
combustion) the gas does n o t . r e a c t c o m p l e t e l y w i t h oxygen u n l e s s
it is i n , or immediately adjacent to, an i n f l a m e d or h i g h
temperature region. In the absence of pilot diesel, a
stoichiometric gas-air mixture does not appear t o r e a c t t o a
s i g n i f i c a n t e x t e n t ; the exhaust gas a n a l y s i s shows no sign of
the presence of carbon d i o x i d e , carbon monoxide, or a l d e h y d e s .
T h e r e f o r e , u n l e s s the n a t u r a l gas i s in a comparatively high
temperature r e g i o n , i t i s u n l i k e l y t h a t the gas would r e a c t when
its c o n c e n t r a t i o n i s lower than the lower l i m i t concentration.
The t e s t s of E l l i o t and D a v i s i n s e v e r a l d i e s e l engines showed
t h a t the lower l i m i t of f l a m m a b i l i t y of n a t u r a l gas i n a i r under
c o n d i t i o n s e x i s t i n g at the end of compression i s a p p r o x i m a t e l y 4
to 5 percent by volume. The corresponding ratio for the
s t o i c h i o m e t r i c m i x t u r e s t r e n g t h was 9.1% by volume.

The experiments by Lewis(l954) were focused mainly on

operations with weak gas-air mixture strength (below 8% by
volume). Lewis used sludge gas(86.6% methane, 5.45% carbon
dioxide, 4.5% n i t r o g e n by volume) i n a s i n g l e c y l i n d e r direct-
i n j e c t i o n engine w i t h compression r a t i o of 14.7:1 and engine
speed of 1000 rpm. The lower l i m i t of f l a m m a b i l i t y i n h i s work
was shown t o be about 6.2% by volume. This i s somewhat higher
than the r e s u l t of e a r l i e r work by E l l i o t and D a v i s ( l 9 5 l ) , and
may be due t o the high c o n c e n t r a t i o n of carbon dioxide and
nitrogen in the s l u d g e gas used by L e w i s . The experiments by
Lewis showed t h a t when the g a s - a i r m i x t u r e s t r e n g t h i s below the
flammability limit, restricting intake air results in
 16

significant increase in the amount of gas reacted and in

increase in i g n i t i o n delay. P r e h e a t i n g of i n t a k e a i r was also
found to increase the amount of gas r e a c t e d , but r e s u l t e d i n
decreased i g n i t i o n delay. W i t h p r e h e a t i n g of the i n t a k e charge
t o 225 deg.C s u b s t a n t i a l l y complete combustion was achieved (95%
of gas r e a c t e d a t 4.5% gas-air mixture s t r e n g t h ,and 70% at 1.0%
mixture strength). The ignition delay of pilot diesel in
r e l a t i o n to g a s - a i r m i x t u r e strength was also studied. The
ignition point was i d e n t i f i e d from the p r e s s u r e - t i m e t r a c e as
the p o i n t of significant pressure rise due to combustion.
F i g u r e 2.1 shows the r e s u l t s o b t a i n e d at c o n s t a n t pilot diesel
r a t e of 0.415 l b / h (0.188 kg/h, 10% of the s t r a i g h t d i e s e l full
load fuel rate). Increase in g a s - a i r mixture strength resulted
initially i n longer i g n i t i o n d e l a y of p i l o t d i e s e l . Beyond the
g a s - a i r mixture s t r e n g t h of about 4% by volume, f u r t h e r i n c r e a s e
in mixture strength showed sharp reduction i n the ignition
delay. The v a r i a t i o n i n i g n i t i o n delay f o r the range of mixture
s t r e n g t h of 0 to 8% by volume was about 3 deg. CA.

The e f f e c t on combustion c h a r a c t e r i s t i c s of g a s - a i r mixture

s t r e n g t h and i n t a k e charge temperature of a motored-engine was
studied by Simonson(1954). H i s e x p e r i m e n t s were performed w i t h
methane i n a d i r e c t - i n j e c t i o n s i n g l e - c y l i n d e r engine f o r intake
charge temperatures ranging from 241 t o 325 deg.C and gas-air
mixture strength ranging from 0 t o 5% by volume. The results
from exhaust a n a l y s i s of c o n s t a n t - s p e e d motored t e s t s r e v e a l e d
that increase in intake charge temperature results in more
f a v o u r a b l e c o n d i t i o n s f o r flame propagation.
 13
speed
<
O 12 1000 r p m

o pilot diesel
LU 0.415 Ib/hr
Q
11
>- gas
< methane
_J 10
LU
Q

7
0 2 3 4 5 6 7 8 (%)
GAS IN ENGINE INTAKE BY VOLUME
(Lewis, 1954)

Figure 2.1 - Effect of G a s - A i r M i x t u r e Strength

on Ignition Delay
 18

Further work by Simonson(1955), with fired-engine operation

included s t u d i e s of t h e e f f e c t s of a i r r e s t r i c t i o n , and changes
i n p i l o t d i e s e l r a t e and i n j e c t i o n t i m i n g . A direct-injection
engine ( t h e same engine as t h e one used by L e w i s ( 1 9 5 4 ) ) w i t h
compression r a t i o of 14.7:1 was used w i t h methane a t the speed
of 1000 rpm. Preheating of i n t a k e charge t o 157 deg.C showed
improvements of 20 t o 30% i n f u e l consumption a t p a r t l o a d s (10-
60 p s i or 70-410 kPa i n brake mean effective pressure) with
pilot diesel rate consisting 8.5% of d i e s e l r a t e of s t r a i g h t
diesel f u l l load operation. With t h e same pilot diesel rate,
advancing of injection timing by 6 deg. C A . resulted in
improvement of f u e l consumption by 10 t o 17% i n the same part
load range. Both i n t a k e a i r r e s t r i c t i o n and i n c r e a s e in pilot
d i e s e l r a t e gave s i g n i f i c a n t improvements i n f u e l consumption.
Maximum c y l i n d e r p r e s s u r e was observed t o i n c r e a s e w i t h advanced
injection timing. Tests at brake mean e f f e c t i v e p r e s s u r e of
115 p s i (793 kPa) r e v e a l e d t h a t advancing t h e injection timing
by 8 deg. C A . increased t h e maximum c y l i n d e r p r e s s u r e from
1000 t o 1250psi (7 t o 8.7 MPa). E a r l y i g n i t i o n and r a p i d rates
of pressure r i s e were r e p o r t e d t o set a l i m i t t o the extent t o
which improved performance can be obtained by advancing the
injection timing. With a pilot diesel injection rate of
0.8 l b / h (0.36 kg/h, 16% of s t r a i g h t d i e s e l f u l l load fuel rate)
at 24 deg BTDC (10 deg advance) combustion was rough even a t the
brake mean e f f e c t i v e p r e s s u r e of 30 p s i (210 k P a ) .

Experiments leading to improvement i n part-load fuel

consumption were done by Moore and M i t c h e l l ( 1 9 5 5 ) . A single-

c y l i n d e r d i r e c t - i n j e c t i o n engine w i t h 4.125 i n (105.4 mm) bore

 19

and 6.00 i n (152 mm) s t r o k e was used a t the speed of 1000 rpm.

Tests c a r r i e d out with sludge gas on the e f f e c t s of a i r

restriction , i n t a k e charge p r e h e a t i n g , and a d v a n c i n g injection

t i m i n g showed r e s u l t s s i m i l a r t o those of the experiments by

Simonson (1 955) . As much as 20% improvement i n f u e l consumpti'on

was r e p o r t e d i n each s e p a r a t e t e s t of the above methods. In

reviewing past work they concluded that r a i s i n g the i n t a k e

temperature i s the o n l y p r a c t i c a l way of extending the lower

l i m i t of f l a m m a b i l i t y .

Work on a large engine was described by M i t c h e l l and

Whitehouse(1955). A four-cylinder direct-injection engine of
10 i n (254 mm) bore and 12 i n (305 mm) s t r o k e w i t h compression
r a t i o of 13.5:1 was used a t the speed of 600 rpm to determine
the optimum conditions for r e l i a b i l i t y and e f f i c i e n c y . The
gaseous f u e l used was s l u d g e gas (87.9% methane, 5.5% carbon
dioxide, 4.4% nitrogen). The p i l o t d i e s e l r a t e was a c o n s t a n t
6% of s t r a i g h t d i e s e l f u l l load f u e l rate. T e s t s on the effect
of a i r restriction showed a c o n s i d e r a b l e improvement i n f u e l
consumption: 46% improvement was reported at t h e brake mean
effective pressure of 20 p s i (140 kPa, 26% of f u l l load). It
was a l s o found t h a t w i t h optimum a i r restriction the exhaust
temperature remained a p p r o x i m a t e l y c o n s t a n t ( w i t h i n ± 50 deg.F
or 30 deg.C) a t a l l l o a d s .

E x p e r i m e n t s made by Felt and Steele(1962) showed that

knock-limited maximum power i s d i r e c t l y r e l a t e d t o the a n t i -

knock q u a l i t y of t h e p r i m a r y fuel. A three-cylinder direct-

injection engine w i t h compression r a t i o of 16.2:1 was used w i t h

 20

pilot diesel rate of 1.65 l b / h (0.74 kg/h, 12.8% of s t r a i g h t

d i e s e l f u l l load f u e l r a t e ) . Lead alkyl anti-knock compounds
were found to be q u i t e e f f e c t i v e i n enhancing the a n t i - k n o c k
q u a l i t y of the primary fuel. A mixture consisting of 95%
propane and 5% tetramethyllead was b l e d i n t o the i n t a k e a i r
stream. With addition of 5.5-6.0 gm of lead per therm
(100,000 Btu or 106,000 kJ) of n a t u r a l gas, i t was p o s s i b l e t o
enhance the maximum power of d u a l - f u e l operation with natural
gas (87.1% methane, 5.1% o t h e r h y d r o c a r b o n s , 7.1% n i t r o g e n ) by
28 percent w i t h o u t knock. The knock was d e s c r i b e d as 'audible
high-frequency combustion loss', which was visible on the
pressure-time trace. The knock was described from the
observation of the shape of the p r e s s u r e - t i m e t r a c e as 'end-gas
knock': the knock a r i s i n g from the a u t o i g n i t i o n of the end-gas
ahead of the flame front. I t i s the type of knock which may
occur i n s p a r k - i g n i t i o n e n g i n e s .

The influences o f f u e l - a i r mixture strength, pilot diesel

rate and intake air temperature on k n o c k - l i m i t e d power were
studied in d e t a i l by Karim et al.(1966/67) with a single-
cylinder direct injection engine with compression r a t i o of
14.2:1 and 97.8% methane as the gaseous fuel. The knock was
observed to be associated w i t h a sharp change i n the running
regime of the engine and accompanied by l o u d l y audible sound.
The typical shape of the pressure diagram is shown in
F i g u r e 2.2. The knock-limited power, which is shown in
F i g u r e 2.3, was observed t o decrease w i t h i n c r e a s e i n the i n t a k e
air temperature and/or p i l o t quantity. I t was found t h a t the
knocking occured o n l y i n a c e r t a i n range of m i x t u r e s t r e n g t h ; i f
 21

- 6 0 - 3 0 T D C 3 0 6 0

(Karim et al.,1966/67)

Figure 2.2 - T y p i c a l P r e s s u r e - T i m e Trace of

Non-knocking and Knocking O p e r a t i o n

w i t h Methane as Gaseous F u e l
Figure 2.3 - V a r i a t i o n of Power Output w i t h the

Overall Mixture Strength for D i f f e r e n t

Intake Temperatures
 23

the engine was o p e r a t e d on e i t h e r s i d e of t h a t range of m i x t u r e

s t r e n g t h , knock c o u l d be a v o i d e d . The r e g i o n of knocking was on
the l e a n s i d e of s t o i c h i o m e t r i c m i x t u r e strength. The e f f e c t of
the c e t a n e number of t h e p i l o t d i e s e l on t h e onset of knock was
found t o be s m a l l .

Karim and Kahn(l968) employed heat r e l e a s e a n a l y s i s i n an

attempt to interpret the combustion processes. From the
analysis with a single-cylinder direct-injection engine and
methane as the gaseous fuel, they concluded that dual-fuel
combustion generally undergoes two d i s t i n c t phases. The f i r s t
is mainly a s s o c i a t e d w i t h t h e consumption of the p i l o t fuel
together with part of the gaseous fuel. The second is
a s s o c i a t e d mainly w i t h the gaseous fuel and depends on i t s
concentration and quality. The heat r e l e a s e a n a l y s i s of very
lean operation supported previous experimental evidences that
poor combustion at low load operation i s due mainly t o the
i n a b i l i t y of the gaseous charge t o supplement effectively the
heat r e l e a s e of t h e f i r s t phase. The a n a l y s i s of the o p e r a t i o n
w i t h knock i n d i c a t e d t h a t t h e knock was m a i n l y associated with
rapid simultaneous burning of t h e p i l o t d i e s e l t o g e t h e r with a
s u b s t a n t i a l f r a c t i o n of t h e gaseous charge.

Study of knock-limited maximum power in relation to

compression ratio and engine speed i s i n c l u d e d i n the review

paper by 0 ' N e a l ( 1 9 8 2 ) . The k n o c k - l i m i t e d power increases very

rapidly as t h e compression r a t i o i s reduced. The t r e n d f o r the

k n o c k - l i m i t e d bmep i s t o i n c r e a s e w i t h engine speed. As engine

speed i n c r e a s e s , l e s s time i s a v a i l a b l e f o r t h e end-gas t o reach

 24

the temperature f o r a u t o i g n i t i o n , and thus the onset of knock i s

suppressed.

In summary, the literature reveals that considerable

r e s e a r c h has been done on d u a l - f u e l o p e r a t i o n w i t h the direct-
injection type of diesel engine. At low l o a d s , combustion
s u f f e r s from weak g a s - a i r m i x t u r e s t r e n g t h r e s u l t i n g i n unburned
gas e s c a p i n g w i t h the exhaust. F i g . 2.4 shows typical thermal
e f f i c i e n c i e s of d u a l - f u e l o p e r a t i o n w i t h d i r e c t - i n j e c t i o n diesel
engines. R e s t r i c t i n g or p r e h e a t i n g intake a i r , increasing p i l o t
diesel flow r a t e , and a d v a n c i n g the i n j e c t i o n t i m i n g have been
shown to be e f f e c t i v e f o r improving the f u e l consumption at low
loads. Maximum power output of h i g h l o a d o p e r a t i o n is limited
by the occurence of knock, which appears t o be of the same kind
as in s p a r k - i g n i t i o n engines. Adding lead a l k y l anti-knock
compounds or c o o l i n g the i n t a k e a i r were shown t o be effective
in improving knock-limited maximum power. The literature,
however, seems s i l e n t on d u a l - f u e l diesel operation with the
prechamber type of d i e s e l engine.
 o

Brake Mean Effective Pressure (kPa)

Figure 2.4 - Typical Thermal Efficiencies of Dual-Fuel and Straight Diesel Operation
 26

CH.III Apparatus and I n s t r u m e n t a t i o n

The arrangement of apparatus and instruments is shown in

F i g . 3.1. The engine was coupled to an electromagnetic
dynamometer. The s i g n a l s from the c y l i n d e r p r e s s u r e transducer
were collected by a NEFF/620 data a q u i s i t i o n u n i t . The data
obtained from various instruments were processed with a
PDP 11/34 computer. The flow diagram of a i r , n a t u r a l gas, and
exhaust gas i s p r o v i d e d i n F i g . 3.2. N a t u r a l gas i s mixed with
air prior to entering the t u r b o c h a r g e r . The g a s - a i r mixture
e x i t i n g from the t u r b o c h a r g e r e n t e r s the cylinder during the
intake stroke. The exhaust gas from the c y l i n d e r passes through
the t u r b i n e s i d e of the t u r b o c h a r g e r , and then e n t e r s a m u f f l e r
b e f o r e d i s c h a r g i n g t o open a i r .

3.1 Engine and Test Bed

Engine

A c a t e r p i l l a r 3304 f o u r - c y l i n d e r engine coupled with an

electromagnetic dynamometer was used in this project. The

engine speed was adjusted to 1600 rpm throughout the

experiments. Full load at this speed was s p e c i f i e d by the

manufacturer as 124 p s i (856 kPa) in brake mean effective

pressure. The engine i s of prechamber type and i s equipped with

a turbocharger. The s p e c i f i c a t i o n of the engine i s provided i n

 diesel
=C3T1
load diesel flow rate^ 1
control •pressure charge
7 transducer amp.

[toothed
dynamometer wheel
CAT. 3 3 0 4 N E F F data
TT acqufeit unit
^-optical
loaa cell

intake pressure^ turbocharger

•I pickup

PDP 11
minicomputer
gas mixer
gas flow rate^A air. . ,.
gas valve —^ restriction t exhaust printer
terminal
air

T flow rate

air
Figure 3.1 - Layout of Apparatus and I n s t r u m e n t a t i o n
 exhaust

F i g u r e 3.2 - Flow Diagram of A i r , F u e l , and Exhaust Gas

 2 9

BORE 1 2 . 1 cm ( 4 . 7 5 in. )

STROKE 1 5 . 2 cm ( 6 . 0 in.)

DISPLACEMENT 6 9 7 0 c m 3
( 4 2 5 cu.in.)
COMPRESSION RATIO 1 7 . 5 : 1

NUMBER OF CYLINDERS 4

MAXIMUM POWER 9 3 . 2 kW ( 1 2 5 hp)

at 2 0 0 0 rpm

TYPE prechamber
ASPIRATION turbo-charged

T a b l e 3 . 1 - Engine Specification
 30

SCALE

2.5:1

A - PRECHAMBER

B - MAIN CHAMBER

Figure 3.3 - S h a p e of Combustion Chambers

 31

Table 3.1. The volume of the prechamber i s 27 percent of t h e

total volume when t h e p i s t o n i s a t the t o p dead c e n t e r . The
cross-section of prechamber and main chamber i s shown i n
F i g . 3.3.

D i e s e l I n j e c t i o n System

The engine was equipped w i t h a s l e e v e m e t e r i n g type of f u e l

system. F i g . 3.4 shows the l a y o u t of the system. The main
components of the s l e e v e metering diesel-injection pump a r e
shown i n F i g . 3.5. The p l u n g e r i s moved up and down i n s i d e t h e
b a r r e l by the a c t i o n of t h e pump camshaft. F i g . 3.6 shows t h e
effective stroke of the p l u n g e r and t h e sequence of t h e
i n j e c t i o n events. The s l e e v e m e t e r i n g system uses centrifugal
governor f l y w e i g h t s (shown i n F i g . 3.7) i n o r d e r t o prevent any
change i n engine speed due t o the v a r i a t i o n i n l o a d . Thus once
the governor control i s s e t a t c e r t a i n p o s i t i o n i n t h e rack
s e t t i n g , t h e engine speed i s maintained constant r e g a r d l e s s of
p o s s i b l e change i n l o a d . F i g . 3.8 shows t h e s i n g l e h o l e d i e s e l -
injection nozzle used i n the e n g i n e . The i n j e c t i o n t i m i n g was
adjusted to desired angle according to the service
manual(NO.SER7053-01, pp80).
 32

F i g u r e 3.4 - S l e e v e M e t e r i n g F u e l System

Roller Follower

F i g u r e 3.5 - F u e l I n j e c t i o n Pump and Housing

 33

SPILL
PORT
PORT

1— H L LINl* J-BEC::. 3 • CO*.7;SJE

EFFECTIVE STROKE INJECTION INJECTION

F i g u r e 3.6 - Sequence of I n j e c t i o n Events

GOVERNOR
BELLCRANK SHAFT
CONTROL SHAFT
• ELLCRANK-
GOVERNOR
CARRIER
CONTROL LEVER

SPRING SEATS

THRUST GOVERNOR
SPRING
COLLAR
GOVERNOR
DRIVE GOVERNOR COVER OF
SHAFT FLYWEIGHTS G O V E R N O R FLYWEIGHTS

F i g u r e 3.7 - Governor Components of S l e e v e M e t e r i n g

Figure 3.8 - Fuel Injection Nozzle
 35

Turbocharger

The engine was equipped w i t h a T1210 model t u r b o c h a r g e r

manufactured by A i R e s e a r c h . A cutaway view of the turbocharger
i s shown i n F i g . 3.9.

Dynamometer

The engine was c o u p l e d t o an e l e c t r o m a g n e t i c dynamometer

(General E l e c t r i c , model 1G136). The a b s o r p t i o n c a p a c i t y of the
dynamometer was 200 horse power (150 kW).

Gas-mixer

In o r d e r t o i n t r o d u c e the n a t u r a l gas, a simple gas-mixer

was installed on upstream of the i n l e t t o the t u r b o c h a r g e r .
F i g . 3.10 shows the gas-mixer.

Intake A i r R e s t r i c t i o n

A s i m p l e ' b u t t e r f l y ' type of v a l v e was i n s t a l l e d near the

gas-mixer t o c o n t r o l the amount of a i r i n t a k e .
Figure 3.9 - Turbocharger Cutaway View
37
 38

3.2 I n s t r u m e n t a t i o n

Torque

The torque a p p l i e d by t h e engine s h a f t t o dynamometer was

o b t a i n e d by p l a c i n g a s t r a i n gage load cell (Interface Inc.,
model 1420-4F) on the dynamometer housing. The maximum
a l l o w a b l e l o a d s p e c i f i e d by t h e manufacturer was 500 l b . A
bridge a m p l i f i e r meter ( E l l i s A s s o c i a t e s , model BAM-1) was used
to a m p l i f y the response from t h e l o a d c e l l . The l o a d cell was
calibrated by placing various weights on the arm of the
dynamometer h o u s i n g and r e a d i n g the v o l t a g e from the b r i d g e
amplifier meter. The calibration curve f o r the l o a d c e l l i s
p r o v i d e d i n Appendix A. The r e l a t i o n between, the response of
the l o a d c e l l and the a p p l i e d weight was very n e a r l y linear.

Cylinder Pressure

The no. 1 cylinder was instrumented w i t h an AVL p i e z o -

e l e c t r i c pressure transducer (model 8QP500c). The transducer

was c o o l e d w i t h water and mounted i n a s t e e l s l e e v e through the

c y l i n d e r head. F i g . 3.11 shows the l o c a t i o n of the mounted

transducer. The s i g n a l from t h e t r a n s d u c e r was t r a n s m i t t e d by a

low n o i s e c a b l e t o a charge a m p l i f i e r ( K i s t l e r , model 5004) and

then t o a d a t a a c q u i s i t i o n system. The system consisted of a

NEFF, System 620, analogue to digital converter which was

connected t o a PDP 11/34 minicomputer. A computer program was

 S C A L E 1.7:1

F i g u r e 3 . 1 1 - Mounting of C y l i n d e r P r e s s u r e Transducer
 40

written (see Appendix C) to sample the pressure s i g n a l at

i n t e r v a l s of one degree crank a n g l e , a l o n g w i t h a bottom dead
center signal, drawn from an optical sensor mounted on the
t o o t h e d wheel at the f r o n t of the engine. The program computed
an ensemble-averaged value of pressure c o l l e c t e d over 30 to
50 c y c l e s f o r each degree of crank a n g l e . The averaged values
were then used t o compute i n d i c a t e d mean e f f e c t i v e p r e s s u r e and
to a n a l y z e apparent energy release. The p r e s s u r e t r a n s d u c e r was
c a l i b r a t e d u s i n g a dead-weight t e s t e r f o r a p r e s s u r e range of 0
to 2000 p s i (14 MPa). The calibration curve i s p r o v i d e d i n
Appendix A.

A i r Flow Rate

A l a m i n a r flow element (Meriam Instrument, model 50MC2-4F,

range 0-400SCFM) was used t o measure the a i r flow r a t e . I t was
mounted between the a i r f i l t e r and t u r b o c h a r g e r . The pressure
d r o p • a c r o s s the element was read i n i n c h e s of water on a water-
f i l l e d U-tube manometer and t r a n s l a t e d to v o l u m e t r i c flow r a t e .
The c a l i b r a t i o n curve p r o v i d e d by the element manufacturer is
shown i n Appendix A.

Gas Flow Rate

The natural gas was drawn i n from the mains s u p p l y at a

p r e s s u r e of 5 p s i . The gas was then passed through a pressure

regulator which reduced the p r e s s u r e t o a p r e s s u r e a few inches

 41

of water h i g h e r than atmospheric pressure. A laminar flow

element (Meriam I n s t r u m e n t , model 50MH10-1.25NT, r a n g e 0-15SCFM)

was mounted t o measure the flow r a t e . The pressure drop across

the element was read i n i n c h e s of water on a w a t e r - f i l l e d U-tube

manometer and t r a n s l a t e d to volumetric flow rate. The flow rate

was then c o r r e c t e d f o r n a t u r a l gas by m u l t i p l y i n g t h e ratio of

v i s c o s i t i e s f o r a i r and n a t u r a l gas. The a m o u n t o f g a s a d m i t t e d

to intake was controlled manually w i t h a t a p e r e d t y p e o f gas

valve. The calibration curve provided by the element

manufacturer i s shown i n Appendix A.

D i e s e l Flow Rate

A p o s i t i v e d i s p l a c e m e n t type of flow m e t e r (American Meter,

model 1A, range 0.05-5GPH) which m e a s u r e s c u m u l a t e d flow rate

was used w i t h a s t o p watch f o r a time p e r i o d o f 10 t o 20 minutes

to o b t a i n the v o l u m e t r i c f l o w r a t e of d i e s e l . The accuracy of
the f l o w meter was c o n f i r m e d by u s i n g a g r a d u a t e d cylinder.

Turbocharger Inlet Pressure

As a p r e c a u t i o n t o p r e v e n t the p o s s i b i l i t y o f t u r b o c h a r g e r

surge due t o e x c e s s i v e a i r r e s t r i c t i o n , the a i r p r e s s u r e at the

inlet of compressor s i d e of the t u r b o c h a r g e r was measured f o r

each change of a i r flow r a t e . A water-filled U-tube manometer

was used. The l i m i t i n g measure of the compressor i n l e t pressure

specified by the engine manufacturer was 24 i n c h e s of water

 42

below the a t m o s p h e r i c p r e s s u r e .

Intake A i r Pressure

In o r d e r t o measure the intake a i r pressure after the

turbocharger, a bourdon-tube type of p r e s s u r e gage (Marquette,
model 41-123, range 30 i n c h e s of water vaccuam t o 15 p s i above
atmospheric pressure) was mounted on the i n t a k e m a n i f o l d near
no. 4 c y l i n d e r .

Engine Speed

The engine speed was measured by a hand d i g i t a l tachometer

(Shimpo, model DT-205) which sends out a c o n t i n u o u s l i g h t beam
and counts the p u l s e s r e f l e c t e d o f f a p i e c e of reflective tape
a t t a c h e d on the engine shaft.
 43

3.3 F u e l

Diesel

The diesel fuel used throughout the experiments had the

following typical properties:
API g r a v i t y - 31

specific gravity - 0.871

lower h e a t i n g v a l u e - 45,263 k j / k g
In the analysis of energy release and computation of
stoichiometric f u e l - a i r r a t i o dodecane (CiaH^) was assumed t o be
the r e p r e s e n t i n g hydrocarbon f o r the d i e s e l fuel.

N a t u r a l Gas

The n a t u r a l gas used i n the experiments had the following

propert i e s :
density - 0.766 kg/m 3

at 101.3 kPa,
25 deg. C

viscosity - 108.96 m i c r o p o i s e

at 21.1 deg. C

lower h e a t i n g v a l u e - 48,558 kJ/kg

A typical c o m p o s i t i o n of the n a t u r a l gas used here i s g i v e n i n

T a b l e 3.2.
 44

COMPOSITION RELATIVE VOLUME (%)

methane 94.00
ethane 3.30

propane 1 .00

i so-butane 0.15
n-butane 0.20
i so-pentane 0.02
n-pentane 0.02
ni trogen 1 .00
carbon d i o x i d e 0.30
hexane 0.01

T a b l e 3.2 - T y p i c a l C o m p o s i t i o n of t h e N a t u r a l Gas Used

 45

3.4 Data Process

The m e a s u r e m e n t s obtained from various instruments for each

experiment were fed into the PDP 11/34 c o m p u t e r for computation

of the following:

. power output

. thermal efficiency

. volumetric efficiency

. proportion of diesel to total fuel input

based on h e a t i n g values

. gas-air, diesel-air, total fuel-air ratio

A typical output from the data process is shown in Table 3.3. A

listing of the computer program used for the computations is

included in Appendix II.

 46

engine speed 1 601 rpm

air flow r a t e 170 c u.f t/m i n

d i e s e l flow r a t e 1 .82 1/hr

gas flow r a t e 6.32 c u.f t/m i n
f u e l consumption r a t e 11,683 BTU/hr-hp
power output 35.6 hp
brake mean e f f e c t i v e p r e s s u r e 4 1.4 ps i
thermal efficiency 21.8 %

volumetric efficiency 86.3 %

d i e s e l input p r o p o r t i o n 15.5 %

gas-air equivalence ratio 0.365

d i e s e l - a i r equivalence ratio 0.068
fuel-air equivalence ratio 0.433
(total fuel)

Table 3.3 - T y p i c a l Output of Computer Program

f o r Data Processing
 47

CH. IV. Experimental Results

4.1 F u e l Consumption

4.1.1 F u e l Consumption w i t h Unmodified Engine

Tests were carried out to study the change in fuel

consumption rate with v a r i a t i o n i n l o a d and p i l o t d i e s e l flow
rate. The engine speed was set at 1600 rpm for all the
experiments. The i n j e c t i o n t i m i n g of the d i e s e l f u e l remained
constant a t 12.3 degrees b e f o r e top dead c e n t r e .

T e s t s always s t a r t e d from s t r a i g h t d i e s e l o p e r a t i o n . While

the engine was running on 100 percent d i e s e l f u e l , the speed was
a d j u s t e d w i t h the d i e s e l f u e l governor rack s e t t i n g t o 1600 rpm,
and the l o a d t o the p r e d e t e r m i n e d s e t t i n g . The gas was then
g r a d u a l l y added by opening the gas c o n t r o l v a l v e . As the amount
of gas was increased, the flow of d i e s e l f u e l was reduced t o
m a i n t a i n the speed at 1600 rpm.

In F i g . 4.1, brake thermal efficiencies based on lower

heating value are p l o t t e d a g a i n s t the f r a c t i o n a l d i e s e l energy

i n p u t , which i s here d e f i n e d as the r a t i o of diesel to total

energy input. The shaded a r e a on the l e f t c o r r e s p o n d s t o the

region where stable operation is not possible because of

insufficient pilot diesel flow rate. In this r e g i o n , the

o p e r a t i o n was erratic with misfired cycles observed on the

cylinder pressure-time trace on the oscilloscope. The

interpolated paths of constant pilot diesel operation are

Brake Thermal Efficiency (%)

8fr
 49

indicated by the dotted lines. As previously described i n

Chapter I I I , the diesel injection rate of the engine was
controlled by two mechanisms - the governor rack s e t t i n g , which
i s a d j u s t e d m a n u a l l y , and the c e n t r i f u g a l governor flyweights,
which respond to any s m a l l change i n a p p l i e d l o a d i n order t o
maintain a constant speed. Thus even though the rack s e t t i n g i s
h e l d f i x e d , a d d i t i o n of gas i n the i n t a k e would r e s u l t i n change
in d i e s e l i n j e c t i o n rate. For t h i s reason i t was not possible
to c a r r y out t e s t s f o r c o n s t a n t p i l o t d i e s e l flow r a t e .

It is seen from F i g . 4.1 that the decrease of thermal

e f f i c i e n c y due t o r e d u c t i o n of of p i l o t d i e s e l flow r a t e becomes
s m a l l e r as the l o a d i n c r e a s e s . At the brake mean effective
pressure of 714 kPa (84 % of f u l l l o a d ) , the change i n thermal
e f f i c i e n c y i s l e s s than 1 percent over the range of fractional
diesel energy input of 7 t o 100 p e r c e n t . The e x t e n t t o which
the f u e l consumption r a t e depends on the p i l o t d i e s e l f l o w rate
for i d l i n g operation i s seen i n F i g . 4.2. The i d l i n g operation
w i t h the f r a c t i o n a l d i e s e l energy i n p u t of 15 percent requires
nearly twice as much energy i n p u t than t h a t of s t r a i g h t d i e s e l
operation.

To o b t a i n another p i c t u r e of t h e r o l e of p i l o t d i e s e l flow

r a t e i n d u a l - f u e l o p e r a t i o n , c a l c u l a t i o n s were made of i n d i c a t e d

thermal efficiencies. The i n d i c a t e d thermal efficiency, being

based on t o t a l power produced (including the power used to

overcome the frictional and pumping l o s s e s ) , s h o u l d be more

closely correlated with the effectiveness of the combustion

p r o c e s s than the engine e f f i c i e n c y based on s h a f t power output.

OS
 51

This assumption i s c o n s i s t e n t w i t h the work of Simonson (1955)

i n which the measured v a l u e s of i n d i c a t e d t h e r m a l e f f i c i e n c y and
the p r o p o r t i o n of gaseous fuel reacted showed qualitatively
s i m i l a r trends. Appendix B i n c l u d e s a d e s c r i p t i o n of the method
used in present work to obtain indicated mean effective
pressure.

The i n d i c a t e d t h e r m a l e f f i c i e n c i e s were interpolated and

plotted in F i g . 4.3. The e f f i c i e n c i e s a r e shown as a f u n c t i o n
of the flow r a t e of p i l o t d i e s e l at constant gas-air mixture
strength. . The equivalence r a t i o 4>g of the gas was computed as
the r a t i o of the mass of the stoichiometric amount of a i r
required for combustion of the gas a l o n e t o the mass of the
a c t u a l amount of a i r drawn i n : i . e .
4>g = s t o i c h i o m e t r i c a i r mass flow rate
a c t u a l a i r mass flow rate

I t can be seen from F i g . 4.3 t h a t the p i l o t d i e s e l flow r a t e has

a very s i g n i f i c a n t e f f e c t on the t h e r m a l e f f i c i e n c y at low gas-
air mixture strength. As the g a s - a i r m i x t u r e s t r e n g t h becomes
r i c h e r , the t h e r m a l e f f i c i e n c y becomes less sensitive to the
change i n p i l o t d i e s e l flow r a t e . W i t h gas e q u i v a l e n c e r a t i o of
0.6, an i n c r e a s e of p i l o t d i e s e l flow r a t e from 1.4 t o 3.5 kg/h
r e s u l t s i n l e s s than 1 p e r c e n t improvement i n i n d i c a t e d thermal
efficiency. If i t i s assumed t h a t the measured v a l u e s of the
i n d i c a t e d thermal e f f i c i e n c i e s are l a r g e l y a function of the
amount of gas burned, then the g a s - a i r e q u i v a l e n c e r a t i o of 0.6
would be a good approximation for the lower limit of
flammability for this particular type of engine. The gas
 53

equivalence r a t i o of 0.6 c o r r e s p o n d s t o the gas-air volumetric

ratio of about 4 percent. T h i s approximate v a l u e i s c l o s e to

the lower i n f l a m m a b i l i t y l i m i t of 4 t o 5 percent suggested by

E l l i o t t and D a v i s ( l 9 5 l ) f o r a d i r e c t - i n j e c t i o n d i e s e l engine.

The d o t t e d l i n e s i n F i g . 4.3 i n d i c a t e the p a t h s of c o n s t a n t

load operation with v a r i a t i o n in p i l o t diesel rate. The l i n e a t
the bottom traces t h e o p e r a t i o n a t i d l i n g , and the one a t the
t o p the o p e r a t i o n a t about 84 percent of f u l l load. Operating
with a constant p i l o t d i e s e l r a t e of 3.4 kg/h would correspond
to 100 p e r c e n t d i e s e l f u e l l i n g a t i d l i n g and 20 p e r c e n t diesel
fuelling at 85 p e r c e n t of full load. The shaded area on the
left i s the r e g i o n of u n s t a b l e o p e r a t i o n due t o m i s f i r i n g .

F i g . 4.4 shows brake thermal efficiency for straight diesel

and d u a l - f u e l o p e r a t i o n s . The s t r a i g h t d i e s e l operation shows
the fuel consumption characteristics typical of compression
i g n i t i o n engines. As t h e l o a d i s increased from idling, the
efficiency improves because of the i n c r e a s e i n power output
w h i l e the f r i c t i o n a l loss of the engine remains relatively
constant for a fixed speed. At the brake mean e f f e c t i v e
pressure of about 700 kPa, the thermal efficiency reaches a
peak. With further increase in load, the curve starts to
d e c l i n e , presumably due t o the decreased access of fuel to
oxygen. One of the d u a l - f u e l e f f i c i e n c y c u r v e s c o r r e s p o n d s t o
o p e r a t i o n w i t h t h e minimum p i l o t diesel flow rate needed to
assure stable operation without misfiring. This curve is
obtained from F i g . 4.1 by i n t e r p o l a t i o n . The t h e r m a l efficiency
of t h e d u a l - f u e l o p e r a t i o n a t p a r t l o a d i s s u b s t a n t i a l l y lower
 o
CD .

Figure 4 .4 - Comparison of Brake Thermal Efficiencies for

Dual-Fuel and Straight Diesel Operation

 55

than that of straight diesel operation. At the brake mean

e f f e c t i v e p r e s s u r e of 850 kPa, which i s about the rated full
load at 1600 rpm, t h e thermal e f f i c i e n c i e s of both operations
are same. E x t r a p o l a t i o n of the dual-fuel operation curve to
higher loads suggests the trend of the thermal efficiency
s u r p a s s i n g t h a t of s t r a i g h t d i e s e l o p e r a t i o n . T h i s t r e n d may be
due t o the homogeneous nature of g a s - a i r mixture in dual-fuel
o p e r a t i o n , which a l l o w s b e t t e r a c c e s s of f u e l t o oxygen.

4 . 1 . 2 E f f e c t of R e s t r i c t i n g I n t a k e A i r

Tests to determined the e f f e c t on thermal e f f i c i e n c y of

r e s t r i c t i n g i n t a k e a i r were performed f o r s e v e r a l l o a d settings
ranging from i d l i n g t o about 50 percent of f u l l load. For each
l o a d s e t t i n g , the r a t e of gas flow was c o n t r o l l e d so that the
flow r a t e of p i l o t d i e s e l would remain a p p r o x i m a t e l y constant.
The a i r restriction was limited by the minimum allowable
p r e s s u r e of the a i r a t the i n l e t of the t u r b o c h a r g e r . Excessive
restriction of a i r beyond the l i m i t r e s u l t e d i n s u r g i n g of the
turbocharger, which i n t u r n caused v i o l e n t u n s t e a d i n e s s of the
engine. Because of t h i s , the a i r - g a s flow r a t i o r e d u c t i o n had
to be c o n f i n e d t o about 10 p e r c e n t . Sufficient reduction of
air, either by using g a t i n g or e l i m i n a t i n g t h e t u r b o c h a r g e r ,
would have i n c r e a s e d the m i x t u r e strength to near the lower
l i m i t of f l a m m a b i l i t y . Such an i n c r e a s e i n the m i x t u r e strength
may have l e d t o much improved f u e l consumption.
 56

F i g . 4.5 shows- t h e e f f e c t of a i r r e s t r i c t i o n on brake

thermal efficiency at various load settings. The gas-air
mixture strength i s r e p r e s e n t e d by t h e e q u i v a l e n c e r a t i o based
on t h e gas a l o n e . The t e s t s f o r a l l t h e l o a d settings exhibit
improvements when a i r resriction i s imposed. The i n c r e a s e i n
brake thermal e f f i c i e n c y however was f o r a l l cases l e s s than 1
percent f o r about 10 percent a i r r e d u c t i o n .

4.1.3 E f f e c t of V a r y i n g I n j e c t i o n Timing

The i g n i t i o n d e l a y of p i l o t d i e s e l was s t u d i e d by o b s e r v i n g
the averaged cylinder pressure vs crank angle trace. In
d e t e r m i n i n g the p o i n t of the s t a r t of ignition (the point at
which combustion has proceeded f a r enough t o a f f e c t the p r e s s u r e
n o t i c e a b l y ) , the e a r l i e s t s i g n i f i c a n t d e v i a t i o n of p r e s s u r e from
the expected compression curve near t o p dead c e n t r e was sought
by close examination. In most cases this method allowed
identification of t h e p o i n t w i t h o u t d i f f i c u l t y . F i g . 4.6 shows
a t y p i c a l pressure trace and identification of the apparent
p o i n t of i g n i t i o n .

With the injection t i m i n g f i x e d a t 12.3 degree BTDC, t h e

change i n i g n i t i o n d e l a y of p i l o t d i e s e l w i t h addition of gas

was investigated at a range of l o a d s e t t i n g s . F i g . 4.7 shows

the crank angle a t t h e s t a r t of i g n i t i o n w i t h t h e pilot diesel

rate varying from 10-20 t o 100 p e r c e n t of t o t a l energy i n p u t .

For s t r a i g h t d i e s e l o p e r a t i o n a t low l o a d , t h e i g n i t i o n does not

 p
CD*

IT)

1.5 kg/h
CD -o
571 kPa
u
QJ IT)
CN

278 kPQ
UJ <=»
CM

—
' I cn
ro

QJ in
b^ep
a
ro —

CD
1/5

o
CO
i—i—i—i—r i—i—r i—r i — i — r
0.2 0.24 0.28 0.32 0.36 0.4 0.44 0.48 0.52 0.56 0.6 0.64 0.6B
Gas-Rir Mixture Strength (equivalence r a t i o based on gasJ

F i g u r e 4.5 - E f f e c t of Intake A i r R e s t r i c t i o n on Brake Thermal E f f i c i e n c y

 58

Figure 4.6 - T y p i c a l C y l i n d e r Pressure Trace and

Apparent Point of I g n i t i o n S t a r t
Figure 4.7 - Apparent Point of Ignition Start at V a r i o u s Loads
 60

s t a r t u n t i l 2 degrees a f t e r t o p dead c e n t r e . As the load is

raised t o h a l f of f u l l l o a d , t h e i g n i t i o n p o i n t i s advanced by a
degree probably due t o t h e i n c r e a s e d end-gas t e m p e r a t u r e . The
pressure vs crank a n g l e t r a c e r e v e a l s the increase i n end-gas
pressure of about 50 kPa f o r t h e above l o a d i n c r e a s e . The
a d d i t i o n of gas extends the i g n i t i o n d e l a y by 1 t o 2 degrees.
The corresponding r e s u l t obtained by Moore and M i t c h e l l (1955)
in a direct i n j e c t i o n engine was about 2 degrees and by Karim
(1980) also in a direct i n j e c t i o n engine about 1.5 degrees.
These agreements i n magnitude of t h e extended ignition delay
seems to suggest that the higher l e v e l of t u r b u l e n c e prior to
the i g n i t i o n of p i l o t d i e s e l i n prechamber engine does not have
a significant effect on i g n i t i o n d e l a y of the p i l o t d i e s e l . In
the absence of a i r r e s t r i c t i o n , t h e p o i n t of ignition i s no
later than 3 degrees ATDC f o r a l l the l o a d s and p i l o t d i e s e l
flow r a t e s t e s t e d .

To study the e f f e c t on thermal efficiency of varying

injection timing, t h e t i m i n g was advanced by 5 and 10 degrees
f o r two l o a d s e t t i n g s , 143 and 278 kPa i n brake mean effective
pressure. The flow r a t e of p i l o t d i e s e l was c o n t r o l l e d t o be
about 20 p e r c e n t of t o t a l energy input. The lower graph of
F i g . 4.8 shows t h a t advancing t h e i n j e c t i o n t i m i n g by 5 degrees
advances t h e p o i n t of i g n i t i o n t o t o p dead c e n t e r f o r both of
the load conditions. The improvement i n brake thermal
efficiency f o r the corresponding change i n i n j e c t i o n timing is
very little, as shown by t h e upper graph. F u r t h e r advance of
the t i m i n g by 5 degrees r e s u l t s i n advancement of the apparent
point of ignition start to 4 degrees BTDC. The thermal
 61

CN

C3
CD . bmep 278 kPa
CM

a CM
OJ

ro

LO
143 kPa
OJ ~~
ro
J- cr, ,

T r i 1—i—r i i i r i—i—i—r

-1—1

V- 1

—I

•Is I
ro -
CZl

c£ T - 1

in.
i
~i i r~i—i—i—i—i—i—i—r
20.0 22.0 24.0 26.0
8.0 10.0 12.0 14.0 16.0 18.0

Injection Timing (deg BTDC)

Figure 4.8 - E f f e c t o f V a r y i n g I n j e c t i o n T i m i n g on Brake Thermal
E f f i c i e n c y and Apparent P o n i t o f I g n i t i o n S t a r t
 62

efficiency o f t h e low l o a d o p e r a t i o n o f 143 kPa in brake mean

effective pressure shows no improvement, while that of the

operation o f 278 kPa i n b r a k e mean e f f e c t i v e p r e s s u r e exhibits a

trend to deteriorate. The t e s t s f o r h i g h e r load conditions were

avoided in fear of excessive increase in maximum cylinder

pressure.
 63

4.2 C y l i n d e r P r e s s u r e

4.2.1 C y l i n d e r pressure i n Unmodified Engine

For each randomly selected cycle c y l i n d e r p r e s s u r e was

measured a t every crank a n g l e . The pressure values at each
angle were than averaged over 30-50 c y c l e s . The engine speed
and the i n j e c t i o n t i m i n g were f i x e d a t 1600 rpm and 12.3 degree
BTDC r e s p e c t i v e l y . F i g . 4.9 shows t h e P-V diagrams for straight
diesel o p e r a t i o n w i t h v a r i o u s l o a d s e t t i n g s r a n g i n g from idling
to full load. The low l o a d o p e r a t i o n s r e v e a l the delay until
a f t e r t h e t o p dead c e n t r e of s i g n i f i c a n t r i s e i n p r e s s u r e due t o
combustion. The maximum c y l i n d e r p r e s s u r e reached a t f u l l load
o p e r a t i o n i s about 7870 kPa (1140 p s i ) .

The -cylinder pressure and volume of compression and

expansion processes in internal combustion engines can be
a p p r o x i m a t e l y r e l a t e d by f o l l o w i n g relationships:

n
P V = const
or
Ln(P) = n Ln(V) + c o n s t

The exponent 'n' would be e x a c t l y t h e r a t i o of s p e c i f i c heats i f

the working f l u i d was an i d e a l gas w i t h c o n s t a n t p r o p e r t i e s , and

the compression or expansion process was adiabatic and

frictionless. For a mixture of real gases at combustion

p r e s s u r e s , the i d e a l gas i s a good a p p r o x i m a t i o n and f o r small

 65

frictional effect, small heat transfer t o the w a l l s , slowly

changing p r o p e r t i e s the above r e l a t i o n s h i p i s v a l i d w i t h nearly
constant value of 'n'. The l n ( P ) - l n ( V ) p l o t t h u s p r o v i d e s an
approximate but c o n v e n i e n t means t o i d e n t i f y the p o i n t s of the
beginning and the end of the combustion. F i g . 4.10 shows the
l n ( P ) - l n ( V ) p l o t of the s t r a i g h t d i e s e l o p e r a t i o n . The figure
indicates the separate prechamber and main-chamber s t a g e s of
combustion through t h e i r e f f e c t s on the pressure development.
At full load, the initial combustion starts near t o p dead
centre. A s h o r t p e r i o d of e x p a n s i o n which i s recognized by a
short straight line near the l n ( v o l u m e r a t i o ) of 0.65. The
subsequent stage of combustion i s i d e n t i f i e d by the d e v i a t i o n of
p r e s s u r e from a s t r a i g h t line. This characteristic is less
distinctive at lower l o a d s . The i d l i n g o p e r a t i o n shows no s i g n
of the 2-stage combustion c h a r a c t e r i s t i c . It seems that the
combustion in the first stage corresponds to the initial
incomplete combustion i n the prechamber. The short period of
•expansion process following the first stage is no doubt
a s s o c i a t e d w i t h the escape from the prechamber of the hot
partially burned diesel-air m i x t u r e and m i x i n g w i t h a i r which
has r e s i d e d i n the main chamber. The second stage combustion
then t a k e s p l a c e i n the main chamber. As the l o a d i s reduced t o
idling, the flow rate of diesel f u e l i s reduced t o a s m a l l
amount and the combustion i n the prechamber is probably nearly
complete, requiring little further combustion in the main
chamber.

Fig. 4.11 compares t y p i c a l i n d i c a t o r diagrams of dual-fuel

and straight diesel operations. The dual-fuel operation

 Log C y L Volume R a t i o (V/VbdcJ

Figure 4.10 Ln P-V Diagram of Straight Diesel Operation

 68

e x h i b i t s a much slimmer P-V diagram, w i t h much of t h e combustion

t a k i n g p l a c e w i t h i n , a narrow range of cylinder volume. The
ln(P)-ln(V) diagrams i n Fig.' 4.12 shows t h e very d i f f e r e n t
combustion c h a r a c t e r i s t i c s of dual-fuel and straight diesel
operations. The ln(P)-ln(V) diagram of dual-fuel operation
shows no o b v i o u s i n t e r m e d i a t e e x p a n s i o n p e r i o d which i s quite
distinctively recognized in straight diesel operation. The
effect i s particularly noticeable at high load where the
combustion d u r a t i o n of d u a l - f u e l o p e r a t i o n i s much s h o r t e r than
t h a t of s t r a i g h t d i e s e l o p e r a t i o n . This would be consistent
with rapid propagation of flame f r o n t s i n the g a s - a i r mixture,
in contrast to t h e slower combustion in straight diesel
operat i o n .

F i g . 4.13 shows t h e change i n maximum c y l i n d e r pressure

w i t h l o a d v a r i a t i o n f o r s t r a i g h t d i e s e l and d u a l - f u e l operation
with p i l o t d i e s e l c o n s i s t i n g 20 percent of t o t a l energy i n p u t .
At high load the maximum cylinder pressure of dual-fuel
operation i s very much higher than t h a t of s t r a i g h t d i e s e l
operation. At brake mean e f f e c t i v e p r e s s u r e of 714 kPa, which
corresponds t o 84 p e r c e n t of f u l l l o a d , the maximum p r e s s u r e of
dual-fuel operation i s about 35 p e r c e n t higher than t h a t of t h e
straight d i e s e l operation.

The maximum cylinder pressures of d u a l - f u e l o p e r a t i o n a t

v a r i o u s l o a d s and p i l o t d i e s e l r a t e s a r e shown i n F i g . 4.14.

F i g 4.15 shows t h e comparison of maximum r a t e of cylinder

pressure rise at various load s e t t i n g s . F i g . 4.16 shows t h e

maximum r a t e of p r e s s u r e r i s e at various p i l o t d i e s e l rates.

 1 1 1—I—I I III 1 1 1—I—I I II 1
30 3 5 1 lO" 1
3 5 1 )Q 4

L o g C y l . Volume R a t i o (V/VbdcJ

Figure 4.12 - C o m p a r i s o n o f L n P-V Diagrams, f o r D u a l - F u e l and S t r a i g h t Diesel Operation

 o

Comparison of Maximum C y l i n d e r P r e s s u r e s for

D u a l - F u e l and S t r a i g h t D i e s e l O p e r a t i o n
Figure 4.14 - Maximum C y l i n d e r P r e s s u r e at V a r i o u s Loads
<
o

co _

10

CL

O o

injection timing
X 12.3 d e g . btdc
o
o ~r T T
100 200 300 400 500 600 700 800 900

Brake Mean Effective Pressure (kPa)

F i g u r e 4.15 - Comparison of Maximum Rate of C y l i n d e r P r e s s u r e R i s e

for D u a l - F u e l and S t r a i g h t D i e s e l Operation
ez
 74

One of t h e most s e r i o u s features of dual-fuel operation

w i t h n a t u r a l i s t h e very large increase i n m e c h a n i c a l l o a d i n g of
the engine, unless the p i l o t d i e s e l i n j e c t i o n i s retarded. In
s e c t i o n 4.2.3, i t w i l l be shown how peak p r e s s u r e s and maximum
rates of pressure r i s e can be g r e a t l y reduced by r e t a r d i n g t h e
injection timing.

4.2.2 E f f e c t of R e s t r i c t i n g I n t a k e A i r

The e f f e c t on maximum cylinder pressure of restricting

intake a i r was considered f o r two l o a d c o n d i t i o n s , 428 kPa
(50 p e r c e n t of f u l l l o a d ) and 571 kPa (67 p e r c e n t of f u l l load)
i n brake mean e f f e c t i v e p r e s s u r e . The flow r a t e of p i l o t d i e s e l
was 15 p e r c e n t of total energy input. F i g . 4.17 shows the
change i n maximum c y l i n d e r p r e s s u r e with r e s t r i c t i o n of intake
air. As mentioned before t h e maximum amount of a l l o w a b l e a i r
restriction was limited by the turbocharger surging
characteristics. Substantial reduction i n maximum cylinder
pressure was achieved without exceeding the throttling
limitation. W i t h maximum a i r r e s t r i c t i o n when t h e l o a d i s below
67 percent of f u l l l o a d t h e maximum c y l i n d e r p r e s s u r e does not
exceed t h a t of s t r a i g h t d i e s e l o p e r a t i o n at f u l l l o a d , which is
about 8000 kPa.

Fig. 4.18 shows t h e change i n pressure just before

combustion with reduction of manifold pressure due to

throttling. The p r e s s u r e j u s t p r i o r t o combustion was o b t a i n e d

Figure 4.17 - Effect of Intake Air Restriction on Maximum Cylinder Pressure
9Z
 77

from t h e c y l i n d e r p r e s s u r e vs crank angle trace as the point

just p r i o r t o t h e s i g n i f i c a n t p r e s s u r e r i s e due t o combustion.
For both of the l o a d c o n d i t i o n s the change i n maximum cylinder
pressure i s of t h e same o r d e r of t h e change i n the p r e s s u r e
p r i o r t o combustion. At t h e brake mean effective pressure of
428 kPa the r e d u c t i o n of a i r f l o w r a t e from 4.9 t o 4.3 m /min 3

r e s u l t e d i n the drop i n the p r e s s u r e prior t o combustion of

400 kPa and i n maximum c y l i n d e r p r e s s u r e of 500 kPa. At the
brake mean effective pressure of 571 kPa the r e d u c t i o n i s
450 kPa f o r t h e p r e s s u r e p r i o r t o combustion and 500 kPa f o r the
maximum cylinder pressure when the a i r flow i s r e s t r i c t e d t o
4.38 from 5.01 m /min. 3
T h i s seems t o suggest t h a t the r e d u c t i o n
of maximum c y l i n d e r p r e s s u r e when t h e i n t a k e a i r is restricted
i s due m a i n l y t o the decreased p r e s s u r e p r i o r t o combustion (due
to r e d u c t i o n i n m a n i f o l d p r e s s u r e ) .

The e f f e c t on t h e maximum r a t e of c y l i n d e r p r e s s u r e r i s e of
intake a i r r e s t r i c t i o n i s shown i n F i g . 4.19. The r e s t r i c t i o n
of i n t a k e a i r seems t o r e s u l t i n h i g h e r maximum r a t e of c y l i n d e r
pressure r i s e . T h i s t r e n d i s b e l i e v e d t o be t h e consequence of
the increased gas-air mixture strength which favours flame
propagation.
Figure 4.19 - Effect of Intake Air Restriction
on Maximum Rate of Cylinder Pressure Rise
 79

4.2.3 E f f e c t o f V a r y i n g I n j e c t i o n Timing

The effect on maximum cylinder pressure of retarding

injection timing was considered for a b r a k e mean e f f e c t i v e

p r e s s u r e o f 571 k P a . The f l o w rate of pilot diesel used was

10 p e r c e n t of t o t a l energy i n p u t . The r e d u c e d maximum cylinder

p r e s s u r e a s a r e s u l t of i n j e c t i o n t i m i n g retardation by 2 and

4 d e g r e e s CA are shown in F i g . 4.20. Retardation of merely

4 d e g r e e s r e d u c e d t h e maximum p r e s s u r e by 1850 k P a . The change

in the pressure prior to combustion f o r the corresponding

retardation was 460 kPa a s shown i n F i g . 4.21.

Fig. 4.22 shows t h e P-V diagram for different injection

timings. As the i n j e c t i o n timing i s retarded the point of the

significant pressure rise i s retarded further from t h e t o p dead

centre e x h i b i t i n g l e s s s t e e p and w i d e r t r a c e . The maximum rate

of p r e s s u r e r i s e was a l s o r e d u c e d a s shown in F i g . 4.20. The

c h a n g e i n t h e r m a l e f f i c i e n c y due t o i n j e c t i o n t i m i n g retardation

was l e s s t h a n 0.5 p e r c e n t .

Test at higher load, a t a b r a k e mean e f f e c t i v e p r e s s u r e o f

714 kPa ( 8 3 % o f f u l l load), with 10 p e r c e n t fractional diesel

energy input and 4 degrees r e t a r d a t i o n showed r e d u c t i o n o f peak

p r e s s u r e f r o m 10.2 t o 7.7 MPa a n d maximum r a t e of p r e s s u r e rise

from 2.1 t o 1.4 MPa/deg. The l o s s i n brake thermal e f f i c i e n c y

due t o the i n j e c t i o n retardation was a b o u t 1 percent.

It i s estimated that safe full-load dual-fuel operation

would r e q u i r e 4 t o 6 degrees of i n j e c t i o n t i m i n g r e t a r d a t i o n and

this would result in loss of thermal efficiency of 1 to

80
o l
on i i i i i i i i i—i—i—i—i—i—i—i—|—i—|—i—|—|—|—i—|—|—i—|—,—r
0.0 0.04 0.08 0.12 0.16 0.2 0.24 0.28 0.32 0.36 0.4 0.44 0.48 0.52 0.56 0.6
C y l i n d e r Volume R a t i o (V/Vbdc)

Figure 4.22 - Effect of Varying Injection Timing on P-V Diagram

 83

2 percent. At l o a d s beyond f u l l l o a d , s i n c e i t i s expected t h a t

the thermal e f f i c i e n c y of d u a l - f u e l o p e r a t i o n w i t h o u t injection
r e t a r d a t i o n would surpass t h a t of s t r a i g h t d i e s e l o p e r a t i o n (see
section 4.1.1), the dual-fuel operation witn sufficient
i n j e c t i o n r e t a r d a t i o n t o assure safe level of peak pressure
would still result in thermal efficiency close t o t h a t of
straight diesel operation. Retarding i n j e c t i o n t i m i n g seems to
be a p r a c t i c a l , and n e c e s s a r y means of e n s u r i n g the s a f e d u a l -
f u e l operation at high load.
 84
CH.V Analysis of Apparent Energy R e l e a s e

5.1 G e n e r a l

One of the most effective means of interpreting the

combustion processes in internal combustion engines is the
estimation of the r a t e of c h e m i c a l energy release, or burning
rate, from the measured p r e s s u r e d i s t r i b u t i o n . The combustion
energy r e l e a s e d a f f e c t s the i n t e r n a l energy of the gas mixtures
inside the cylinder, the heat transfer through the c y l i n d e r
w a l l s , and the work done on the piston head. By appropriate
estimation of heat t r a n s f e r and thermodynamic properties of gas
m i x t u r e s , the apparent energy r e l e a s e due t o combustion can be
e s t i m a t e d from measured v a l u e s of c y l i n d e r p r e s s u r e and change i n
c y l i n d e r volume. The a n a l y s i s p r o v i d e s a q u a l i t a t i v e picture of
the combustion processes of dual-fuel and straight diesel
operation.
 85
5.2 Method of c a l c u l a t i o n

5.2.1 D e f i n i t i o n s , E q u a t i o n s , and Assumptions

When both the intake and exhaust v a l v e s a r e c l o s e d the

m i x t u r e s of a i r and burned and unburned f u e l s can be considered
as a system undergoing a change of s t a t e , which i s bounded by
c y l i n d e r w a l l s and p i s t o n head (see F i g . 5.1). The f i r s t law of
thermodynamics then can be a p p l i e d t o the system f o r a s m a l l time
change 6 t :

f i r s t law 6Q = dE + 5W

where
Q - heat t r a n s f e r t o the system
E energy of the system
W - work done by the system

S i n c e the o n l y s i g n i f i c a n t e n e r g i e s of the system involved here

are the internal and c h e m i c a l energy, the energy of the system
can be assumed t o c o n s i s t of the f o l l o w i n g :

E = U + CE
where
U - i n t e r n a l energy of the system
CE - c h e m i c a l energy of the system
Figure 5.1 - Control Volume f o r Apparent Energy Release Analysis
 87

Then t h e c o r r e s p o n d i n g f i r s t law becomes:

6Q = dU +dCE + 6W

For a finite change of t i m e , A t , w i t h the system undergoing a

change from s t a t e / to state the first law can be
integrated to y i e l d the f o l l o w i n g :

•Q.=
1^1+1
AU + ACE + 1.W.i+1.
where
AU = U i + 1 - U.
ACE = CE.., " CE.
1+1 l

I f P i s d e f i n e d t o be the average p r e s s u r e of t h e system :

l i+1 I i+1

then t h e work done on t h e system can be approximated a s :

i•Wi.+ 1 - i• iP •
+ 1, ,i +(V.
1 . - IV. )
J 1

where - c y l i n d e r volume a t i ^ s t a t e
f c

The change i n c h e m i c a l energy of t h e system can be e s t i m a t e d a s :

ACE jm f j U c j
 88

where

m^j - mass o f j t n
fuel burned during

i f c
^ state, (j=1 for diesel,

j=2 for natural gas)

u j c - internal energy of combustion

of j t h
fuel

Now a finite difference form of t h e f i r s t law may be w r i t t e n as:

f fj cj
m U
- iQ l i + " A U
" i P
i + 1 A V
(Eqn 5.1)

chemical heat internal work

energy transfer energy
change change

Evaluation of these terms will be discussed in subsequent

sections.

Assumpt i o n s

I n d e v e l o p i n g t h e method o f c o m p u t a t i o n s e v e r a l assumptions

were made, namely:

1. The constituents of the mixture i n t h e c y l i n d e r behave as

ideal gases with temperature-dependent thermodynamic

properties.

2. The gaseous c o n s t i t u e n t s of t h e m i x t u r e a r e c o n s i d e r e d t o be

homogeneous and u n i f o r m i n thermodynamic state : spatial

non-uniformity in the rate of c h e m i c a l energy r e l e a s e is

ignored.
 89
3. The c o m p o s i t i o n of the combustion products corresponds to
equilibrium dissociation.

4. The continuous variation of thermodynamic p r o p e r t i e s w i t h

time can be a d e q u a t e l y represented by stepwise variation
over a s m a l l time i n t e r v a l c o r r e s p o n d i n g t o 1 degree crank
angle.

5. Presence of r e s i d u a l gas d u r i n g the i n t a k e i s n e g l e c t e d .

6. O v e r l a p of i n t a k e and exhaust v a l v e i s i g n o r e d .
7. At any g i v e n i n s t a n t the burned f r a c t i o n s of the n a t u r a l gas
and d i e s e l f u e l s a r e the same.

Heat T r a n s f e r

In order to account f o r t h e heat t r a n s f e r between the gas

m i x t u r e and t h e c y l i n d e r w a l l s , both convective and radiative
modes were considered, f o l l o w i n g t h e procedure of Annand(l963)
whose e q u a t i o n i s :

q/A = a(k/D)(R) (T - T
b
w a l l > + c(T« - T ^ ^ ) -
where
q - heat t r a n s f e r rate
A - s u r f a c e a r e a of c y l i n d e r walls
k - t h e r m a l c o n d u c t i v i t y of t h e m i x t u r e
D - bore
R - Reynolds number d e f i n e d as pVD//x
where

p - d e n s i t y of t h e m i x t u r e
 90
"V - mean p i s t o n velocity

n - v i s c o s i t y of the m i x t u r e
T - m i x t u r e temperature
T
wall ~ y^^ ^
c n c e r
wall temperature
a ,b,c - c o n s t a n t s

The first term w i t h the f i r s t o r d e r of temperature a c c o u n t s f o r

the c o n v e c t i v e heat t r a n s f e r and t h e second term w i t h t h e fourth
order temperature f o r t h e r a d i a t i v e . The c o n s t a n t s 'a' and 'b'
for the convective term were selected to y i e l d a f i t with
least-square-errors to t h e apparent heat t r a n s f e r d u r i n g the
compression s t r o k e . A n o n l i n e a r l e a s t - s q u a r e s - f i t technique was
adopted i n optimizing t h e two constants f o r a l a r g e s e t of
apparent heat t r a n s f e r data obtained (from equation 5.1 and
measured pressures w i t h no combustion and known c o n s t i t u e n t s of
the m i x t u r e ) f o r d u a l - f u e l and s t r a i g h t d i e s e l o p e r a t i o n over a
range of loads. F i g . 5.2 shows t h e f i t t e d c u r v e and data f o r
s t r a i g h t d i e s e l o p e r a t i o n a t brake mean effective pressure of
571 kPa. The o p t i m i z e d v a l u e s f o r t h e d i m e n s i o n l e s s c o n s t a n t s 'a'
and 'b' were 0.47 and 0.7 r e s p e c t i v e l y . The v a l u e s suggested by
Annand(l963) f o r 'a' was 0.35-0.8 and f o r 'b' was 0.7.

The c o n s t a n t 'c' f o r t h e r a d i a t i o n term was taken t o be

3.3 x 1 0 " 1 1
kJ/K* as suggested by Annand f o r d i e s e l e n g i n e s . T h i s
v a l u e would c o r r e s p o n d t o t h e p r o d u c t of the Stefan-Boltzmann
constant a and an e m i s s i v i t y of 0.58, a p p r o p r i a t e t o grey body
radiation.
 Heat Transfer (kJ/sec)
50 0 -30 0 -10.0 10.0 3 0 . 0 5 0 . 0 7 0 . 0 . 9 0 . 0 110.0 130.0

CO
1 I I I I I I I I I I I ! I I I 1 I—
CD
X I

— I'
I X
* I x
CD x I 3

*l X

cn X
IX
X . ^
X 1

l x

X X

cn X I

i
* i X

v.
a' o
CJ
< X
~t ~I
CD x #
QJ , X ^
\
X V \
te

CO
X ) CD
m
CL
im
— ro
o

cn
in cr 10 01
CD •o 3 Q
a
TJ
T3
CL Q. "CJ <—r
ght CMc

ro
o a
i CL TJ
cn cn
~o
3
o 3
a
55 o 10 o o
5T CL
IJ
Q
3
operation

CD a

ro

cn
-to.
CD

16
 92
It was assumed t h a t the temperature of t h e w a l l , a t a g i v e n
l o a d , s t a y s c o n s t a n t throughout t h e c y c l e , and varies linearly
with the a p p l i e d l o a d . The measurements of w a l l temperature made
by Kamel and Watson(l979) on an i n d i r e c t - i n j e c t i o n R i c a r d o swirl
engine showed t h a t t h e change i n w a l l temperature throughout the
cycle at full load was less than 10 percent of the mean
temperature. Their data a l s o suggested t h a t t h e w a l l temperature
of both prechamber and main chamber v a r i e d n e a r l y linearly with
applied load. In the present work t h e w a l l temperature was
calculated from

T
w a l l " 0.071(bmep) + 540
bmep i n kPa
T ,, i n K
wall

The numbers o b t a i n e d from t h i s f o r m u l a f o r T ^ are well within

10 percent of those measured by Kamel and Watson a t the same
engine speed.

F i g . 5.3 shows a t y p i c a l c a l c u l a t i o n of t h e apparent r a t e of

energy r e l e a s e w i t h and w i t h o u t t h e adopted heat t r a n s f e r model;
the computation procedure i s p r e s e n t e d l a t e r .

Dissociation

In computing the constituents of t h e combustion p r o d u c t s

e q u i l i b r i u m d i s s o c i a t i o n was assumed. The d i s s o c i a t i o n reactions

c o n s i d e r e d a r e as f o l l o w s :
 with heat transfer model
without heat transfer model

OJ straight diesel operation

TJ,
bmep - 571 kPa
speed - 1600 rpm

QJ
in
ro
UJ cn
QJ o '

£"\
QJ^"
•f-»
ro
CH cn
o

v. V , v.

1 1— T
-la.o tdc I0.O 50.0 7D:0
30. D
(deg flTDC)
Crank Angle

Figure 5.3 - Effect o f Heat T r a n s f e r Model on A p p a r e n t Rate o f Energy Release

 94
a. C 0 2 < > CO + 1/2 0 2

b. H 0 < 2 > 1/2 H 2 + OH

c. H 0 < 2 > H 2 + 1/2 0 2

d. 1/2 N 2 + 1/2 0 2 < > NO

It may be noted that i n h i s engine mixture dissociation

calculations, Campbell(1977) c o n s i d e r e d the above d i s s o c i a t i o n s ,
and i n a d d i t i o n t h e d i s s o c i a t i o n of 0 , H , and OH. 2 2 As will be
shown t h e degree of d i s s o c i a t i o n i s small at the r e l a t i v e l y high
p r e s s u r e s and low temperatures of compression ignition engines.
Hence only 4 dissociation processes were considered i n the
equilibrium calculation.

For each s t e p of i n c r e m e n t a l time At and g i v e n p r e s s u r e and

temperature of the m i x t u r e , the following s e t of nonlinear
equations were solved f o r the number of moles of combustion
products:

CO 0, = K.(P'/P)
v
Y
co 2 or
(N +A)[N c o Q +1/2(A+C-D)] 2

2
(\ y ot
H
" VPVP)^
< co N
2 " A
)

YX ' 2 Y
H ? OH
i = K (P°/P)'
R
2

Y
Y
H 0 2

or
u
(N H +1/2B+C) ( N + B ) 2
QH

( tot)""
N 2
- K (P°/P)
B
:

( H 0- - ^N
2
B C
 95

Y Y
H
2 °2
K (P7P)
C
1

Y
H 0 2

o r

(N H +l/2B+C)[(N Q +l/2(A+C-D) ] " 2

1 :
" ^ t o t y k = K [P°/P)
C
:

( H 0" "
N
2
B C )

Y
D
NO _ V

N
2 °2 or

< NO+ )
N D

(NXI - l / 2 D ) [ N 4
N + l/2(A+C-D)] J 5

N
tot = I N
i +
1/2(A + B + C)

where K^,Kg,K^,K^ are equilibrium constants

for reactions a,b,c,d.

^CO'^NO' E T C a r e e c
l i l i
u D r
i- u m

compositions.

Nfjo'^NO' e t C a r e t
^ i e n u m
^ e r
°^ m 0
^ e s

of constituents present prior to the

current step of combustion.
A,B,C,%Dare numbers of moles of CO,,
H 0,H 0,K'
2 2 2 or 0 2 dissociated in
reactions a,b,c,d, respectively.

P° is atmospheric pressure, 101.3 kPa.

P is the gas mixture pressure.

«. i s t h e t o t a l number of moles of
tot
the m i x t u r e .
 96

Given the pressure and temperature of the mixture, the five

equations were simultaneously solved f o r A, B, C, D, and N

t o t

using a modified Newton's method. The values of the equilibrium

constants, which are f u n c t i o n s of temperature, were calculated

from fitted curves based on thermodynamic data given i n the JANAF

Thermodynamics Tables.

Fig. 5.4 shows a typical calculation of the apparent rate of

energy r e l e a s e computed with and without dissociation.

Figure 5.4 - Effect of Equilibrium Dissociation Calculation on Apparent Rate o f Energy Release
 98
5,2.2 Computation Procedure

The apparent energy release is obtained by solving the

equations of mass and energy conservation for the mixture
temperature and the f r a c t i o n of f u e l burned. The work done on the
piston i s computed from the smoothed p r e s s u r e d a t a and change i n
c y l i n d e r volume. The rate of heat transfer is estimated as
described in section 5.2.1. The c o m p o s i t i o n of the m i x t u r e i s
computed w i t h the e q u i l i b r i u m d i s s o c i a t i o n assumption, and with
the e q u a t i o n s p r o v i d e d i n the p r e v i o u s s e c t i o n .

C o n s i d e r a s m a l l s t e p i n the c a l c u l a t i o n during which the

state of the mixture changes from s t a t e / t o s t a t e The
c a l c u l a t i o n s f o r s t a t e (i+1) start with complete knowledge of
s t a t e /; the f o l l o w i n g c o n d i t i o n s are g i v e n :

T
i ' P
i ' CH. i'
( n ) (
C n
1 2 H 2> i6 j h { n

where the n^'s a r e the numbers of moles of CH ,

4

C, H , N ,
2 2 6 2 0 , 2 H 0,2 C0 , 2 CO, H, 2 OH, NO.

The c a l c u l a t i o n procedure i s as follows:

1. Assume T.. 1f (fr). . ( f r a c t i o n of f u e l burned i n one

1+1 I to 1+1
step).
2. O b t a i n the c o m p o s i t i o n molar f r a c t i o n s ( j ) ^ n
+ 1 which would
e x i s t i n s t a t e (i+1) were t h e r e no d i s s o c i a t i o n .
 99

With the assumed T^ + 1 and the measured p r e s s u r e P ^ +1 perform

equilibrium dissociation calculations to obtain the values
of ^ j^i+i
n
which s a t i s f y the d i s s o c i a t i o n r e l a t i o n s h i p s i n
s e c t ion 5.2.1.

For each s p e c i e s c a l c u l a t e the change i n the number of moles

An.= (n.).., -(n.).

3 3 1+1 j 1

Compute the changes i n chemical energy ACE and i n t e r n a l

energy AU as

ACE = ?An . ( u . • + Au .) 0

j 3 3 3 l

where
u°£j = i n t e r n a l energy of f o r m a t i o n

at 298 K
AUj = U j ( T i + 1 ) - Uj(298K)

AU = ? ( n . ) . ( u . ( T . , . ) - u.(T.))
j 3 3 > 3 1 1+ 1

where U j - i n t e r n a l energy of j * " * c o n s t i t u e n t

1

of gas m i x t u r e
6. Compute ^Q^ +1 and i j w
+ 1

7. Check the f o l l o w i n g two c o n s e r v a t i o n e q u a t i o n s :

ACE = .Q. +1 + AU + .W. +1

P
i + 1 i
V
+ 1 = ^ ^ j ^ - H I > R T
i
+ 1

8. I f the above two e q u a t i o n s a r e s a t i s f i e d then computation i s

completed. I f n o t , repeat from 1.
 100

C y l i n d e r P r e s s u r e Data

The c y l i n d e r p r e s s u r e was r e c o r d e d at every degree of crank

a n g l e . The measured v a l u e s were then averaged over 30-50 randomly
selected c y c l e s . I t r e q u i r e d 20-30 minutes t o o b t a i n an averaged
p r e s s u r e t r a c e of 30 c y c l e s w i t h the NEFF data aquisition unit
and the PDP/11 computer. Because of large cycle-to-cycle
v a r i a t i o n s , the r e s u l t i n g p r e s s u r e - c r a n k angle curves were not
smooth enough to provide a smooth c a l c u l a t e d c u r v e of r a t e of
energy r e l e a s e . The p r e s s u r e - c r a n k a n g l e c u r v e s were smoothed by
f i t t i n g a c u r v e between the o b t a i n e d d a t a . The technique involved
f i t t i n g a piece-wise cubic polynomial (continuous to the second
derivative) between the p r e s s u r e measurements over a crank angle
range of 180 degrees w i t h minimization in square errors. The
inverse of the v a r i a t i o n i n the s l o p e of the p r e s s u r e t r a c e f o r
f o u r n e i g h b o u r i n g p o i n t s were used i n p r o v i d i n g the weight for
the l e a s t - s q u a r e s f i t .

Fig. 5.5 shows the r a t e of energy release calculated from

unsmoothed and smoothed p r e s s u r e - c r a n k angle c u r v e s . The smoothed
p r e s s u r e c u r v e showed very small visually detectable change,
except i n the r e g i o n near the peak of the combustion pressure.

Computer Program f o r Apparent Energy Release

The main f u n c t i o n of the computer program was t o execute the

computation procedure d e s c r i b e d in the previous section. The

program initially read in the c y l i n d e r p r e s s u r e d a t a and flow

 smoothed
not smoothed
cn
_ cd"
CD
OJ straight diesel operation
-o
CD bmep - 571 kPa
speed - 1600 r p m

OJ
cn
ro "
OJ
—H (O
QJ o '

"1_

UJ

0 J d

ro
CC
C3 "

CD"

1
-1D.0 tdc 50.0 70.0
10.0 30.0
Crank Angle (deg.RTDC)

Figure 5.5 - E f f e c t o f Smoothing Pressure Data on A p p a r e n t Rate o f Energy Release

 102

r a t e s of a i r and f u e l s a l o n g w i t h d a t a f o r o p e r a t i n g conditions.
The p r e s s u r e data were then smoothed-and volumes of the c y l i n d e r
f o r a l l crank a n g l e s were computed. S u b s e q u e n t l y , f o r each degree
of crank a n g l e , e q u a t i o n s f o r energy and mass c o n s e r v a t i o n were
s i m u l t a n e o u s l y and i t e r a t i v e l y s o l v e d f o r the temperature and the
fraction of f u e l burned. A m o d i f i e d Newton's method was used i n
s o l v i n g the system of n o n l i n e a r e q u a t i o n s . The program assumed
that combustion may take p l a c e at anytime a f t e r the i n j e c t i o n of
d i e s e l f u e l . B e f o r e the d i e s e l i n j e c t i o n p o i n t , the program took
an alternate route and merely computed the m i x t u r e temperature
d i r e c t l y from i d e a l gas law. F i g . 5.6 shows the f l o w c h a r t of the
procedures adopted in the computer program. A l i s t i n g of the
computer program i s p r o v i d e d i n Appendix E. Fig. 5.7 shows a
t y p i c a l output of the program.

Check of Computation

In order to c o n f i r m q u a n t i t a t i v e l y the c o r r e c t n e s s of the

method used, the computed amount of consumed fuel energy per
averaged c y c l e from the computer output was compared w i t h a c t u a l
amount of fuel energy input. Because the computed rate of
chemical energy r e l e a s e i s e s s e n t i a l l y z e r o except f o r the crank
a n g l e i n t e r v a l of -10 t o +90 degrees a f t e r top dead center, it
was n e c e s s a r y t o i n t e g r a t e the energy r e l e a s e o n l y i n t h i s range.

Table 5.1 shows the r a t i o s of the computed t o a c t u a l energy

consumed f o r v a r i o u s o p e r a t i o n s . From a l l the r a t i o s shown i n the

t a b l e , i t i s seen t h a t the agreements between the computed and

 103

read i n
flow r a t e of d i e s e l , gas, a i r
P ( 0 ) , C.A. a t i g n i t i o n , injection

read i n
properties of d i e s e l , gas, a i r ,
combustion products

comp u t e
V(0)

smooth
P(O)
repeat f o r 0 = -89 t o 9JJ

update
number of moles of reactants

compute guess
T from amount of f u e l burnt, T
i d e a l gas law
compute
compute stoichiometric combustion
heat transfer products

set account
energy r e l e a s e = 0 for dissociation

compute
print heat transfer
P, T, h e a t transfer,
energy release

are
set
mass and
0 = 0 + 1 yes energy conserved?

Figure 5.6 - F l o w c h a r t o f Computer Program f o r Apparent

Energy Release Analysis
 104

Figure 5.7- Typical Output of Computer Program for Apparent

Energy Release Analysis
LOAD MODE FRACTIONAL ACTUAL COMPUTED RATIO OF
OF D I E S E L ENERGY ENERGY ENERGY COMPUTED TO A C T U A L
(kPa)
OPERATION INPUT (%) CONSUMED (kJ) CONSUMED (kJ) ENERGY CONSUMED

0 s traight - 0.79 0.83 1.06

diesel

279 s t raigh t - 1 . 74 1.67 0.96

diesel

571 s traight - 3 .05 2.84 0.93

diesel

713 s traight - 3 . 75 3.48 0.93

diesel

856 s traight - 4 . 58 4.26 0.93

diesel

571 dual-fuel 20 . 9 3.17 3.01 0.95

571 dual-fuel 10 . 2 3.14 2 .95 0.94

Table 5.1 - Comparison of A c t u a l and Computed Fuel Energy Consumed

 106

actual energy consumed a r e q u i t e good. The s m a l l disagreements

are p r o b a b l y due t o the v a r i a t i o n i n heating values of fuels,
inadequacy of heat transfer model, unburned f u e l , and v a r i o u s
assumptions made d u r i n g the course of method development. It
should be noted that the amount of unburned gas e s c a p i n g the
c y l i n d e r was not s u b t r a c t e d from the a c t u a l amount of f u e l energy
input.
 107

5.3 Analysis

5.3.1 O p e r a t i o n s w i t h U n m o d i f i e d Engine

The c a l c u l a t e d r a t e of energy r e l e a s e f o r straight diesel

operation with l o a d s r a n g i n g from i d l i n g t o f u l l l o a d i s shown
i n F i g . 5.8. At i d l i n g near the top dead center the curve
declines t o n e g a t i v e v a l u e s p r i o r t o the peak. T h i s may be the
r e s u l t of the inadequacy of the heat transfer model. At low
loads the model seems to underestimate the rate of heat
t r a n s f e r , r e s u l t i n g i n n e g a t i v e v a l u e s f o r the apparent r a t e of
energy release. The roughness of the c u r v e s d u r i n g and towards
the end of combustion i s due t o roughness i n the p r e s s u r e d a t a .
The roughness c o u l d have been reduced by f u r t h e r smoothing the
pressure data.

At full load(bmep = 856 kPa) the r a t e of energy release

r e v e a l s two s t a g e s of c o m b u s t i o n . The f i r s t s t a g e lasts until
12 to 14 degrees a f t e r t o p dead c e n t r e . As the l o a d d e c r e a s e s
the second stage combustion becomes l e s s d i s t i n c t i v e . It seems
that the f i r s t peak c o r r e s p o n d s t o combustion i n the prechamber
and the second to combustion in the main chamber. The
calculated cumulative energy release in F i g . 5.9 further
s u p p o r t s t h i s view. Near 13 degree C.A. after the top dead
centre, which is approximately the p o i n t s e p a r a t i n g the two
s t a g e s of combustion, the cumulative energy release for the
loads o t h e r than i d l e i s n e a r l y the same at about 1.3 k J . This
i s about 30% of the t o t a l cumulative energy release at full
load. The v o l u m e t r i c r a t i o of prechamber t o the t o t a l volume at
Figure 5.8 Rate o f Energy Release of Straight Diesel Operstion at Various Loads
Figure 5.9 Cumulative Energy Release of Straight Diesel Operation at Various Loads
 110

the t o p dead centre i s about 25 p e r c e n t . Thus the a n a l y s i s

s u g g e s t s t h a t t h e combustion i n s t r a i g h t diesel operation of
prechamber engine consists of two distinct and subsequent
stages-prechamber and main chamber. From F i g . 5.9 i t can be
seen that t h e time period d u r i n g which the combustion t a k e s
place increases with the increase i n load. The maximum r a t e of
energy r e l e a s e i s t h e s m a l l e s t a t f u l l l o a d and i n c r e a s e s as t h e
load i s decreased. At v e r y low l o a d s the r a t e of energy r e l e a s e
rises very r a p i d l y a l t h o u g h t h e maximum r a t e of energy r e l e a s e
i s l i m i t e d by t h e t o t a l d i e s e l energy input. F i g . 5.10 shows
the dependence on a i r - f u e l r a t i o of the maximum r a t e of energy
release. I t i s seen t h a t t h e r a p i d i t y of t h e combustion in
straight diesel operation increases nearly l i n e a r l y with a i r -
fuel ratio. •*

F i g . 5.11 shows t h e r a t e of energy r e l e a s e with dual-fuel

operation i n which pilot diesel fuel accounts f o r about
20 p e r c e n t of t h e t o t a l energy i n p u t , and when no change has
been made in diesel injection timing. The most noteworthy
f e a t u r e i s t h e n e a r l y t w o f o l d i n c r e a s e i n t h e maximum rate of
energy release. I t i s observed from the f i g u r e t h a t t h e change
i n combustion d u r a t i o n w i t h v a r y i n g load i s relatively small
when compared t o t h a t of s t r a i g h t d i e s e l o p e r a t i o n . The shapes
of t h e r a t e of energy r e l e a s e c u r v e s e x h i b i t v e r y l i t t l e of t h e
two-staged combustion characteristic which i s observed in
straight d i e s e l operation. The r a p i d i t y of r i s e of t h e r a t e of
energy release increases with the increase i n load. F i g . 5.12
shows t h e maximum r a t e of energy release plotted against the
gas-air mixture strength. The figure •indicates strong
Ill
 n 1 1 1 1 1 1 1 r~

- mn tdc 10.0 30.0 50.0 70.0

Crank Rngle (derj PTDC1

Figure 5.11 - Rate o f Energy Release o f Dual-Fuel Operation at Various Loads

^ O

r _ actual gas-air mass ratio

g stoichiometric gas-air mass ratio

Figure 5.12 - Effect of Gas-Air Mixture S t r e n g t h o n Maximum Rate o f Energy Release

in Dual-Fuel Operation
i
 114

dependence of t h e r a p i d i t y of combustion on t h e g a s - a i r mixture

strength. The c u m u l a t i v e energy r e l e a s e of d u a l - f u e l operation

c o r r e s p o n d i n g t o t h e c u r v e s i n F i g . 5.11 i s shown i n F i g . 5.13.

F i g . 5.14 and F i g . 5.15 compare s t r a i g h t d i e s e l and dual-

fuel operation at 85 and 33 percent of• f u l l load. From
F i g . 5.14 i t i s seen t h a t f o r h i g h l o a d o p e r a t i o n t h e combustion
d u r a t i o n of t h e d u a l - f u e l o p e r a t i o n i s much s h o r t e r than t h a t of
the s t r a i g h t d i e s e l o p e r a t i o n , and the maximum rate of energy
r e l e a s e i s about 3.7 times h i g h e r . The remarkable d i f f e r e n c e i n
the shape of t h e r a t e of energy r e l e a s e f o r t h e two modes of
operations s u p p o r t s the view mentioned i n s e c t i o n 4.2.1 t h a t t h e
mechanisms of combustion are different. Evidently the
combustion i n d u a l - f u e l o p e r a t i o n i s m a i n l y c a r r i e d out by flame
p r o p a g a t i o n through' the premixed g a s - a i r m i x t u r e , i g n i t e d by t h e
burning d i e s e l spray which p e n e t r a t e s t h e main chamber as a hot
jet. In s t r a i g h t d i e s e l o p e r a t i o n , combustion occurs i n the
fuel-air mixture adjacent t o evaporating f u e l drops r a t h e r than
p r o p a g a t i n g as a t u r b u l e n t flame through the e n t i r e mixture.
Since most of the f u e l does not r e q u i r e e v a p o r a t i o n i n the
former case i t i s r e a s o n a b l e t h a t t h i s mode of combustion should
be f a s t e r .

F i g . 5.16 shows the r a t e of energy r e l e a s e f o r a low load

operation (brake mean effective pressure of 279 kPa) with

v a r i o u s flow r a t e s of pilot diesel. The apparent ignition

points, which may be i d e n t i f i e d as t h e p o i n t s where t h e c u r v e s

s t a r t r i s i n g , agree q u i t e w e l l with those observed from the

pressure-crank angle t r a c e . The i g n i t i o n d e l a y i s i n c r e a s e d as

9TI
 LP Straight diesel operation dual-fuel operation
bmep - 279kPa bmep-279kPa
TO 714 kPa (21% diesel)
. T
Ol o 714kPa
QJ
"a (18% diesel)
^ to

QJ
LO c-)
ro
QJ
CO
QJ
cr
Oi CO

LU o*

QJ —
t- a
1

ro

co
o

- < ^ ^ . ^ - ^ V - ~T C7>^p^
a

-I 1 1
1—
-IU.O tdc I0.Q 30.0 50.0
Crank Rngle fdeg flTDCJ

Figure 5.14 - C o m p a r i s o n o f R a t e o f E n e r g y R e l e a s e f o r Straight Diesel and Dual-Fuel

Operation
Figure 5.15 - Comparison of Cumulative Energy Release for Straight Diesel and Dual-Fuel
Operation
 to

CD bmep - 279kPa ( 3 3 % full load)

oo
. -=r
cn CD 13.5 V. diesel
QJ
20.6% diesel
\ rr, 32.0% diesel
100% diesel

O J D_
ir, C
R-)

•o ,1

I I I I I I I I 1

-10 0 tdc 10
-° 30.0 50
-° 7 0 0

Crank Angle (deg RTDC)

Figure 5.16 - Rate o f Energy Release of Dual-Fuel Operation at Various Pilot D i e s e l Flow Rates
 119

much as 2 degree C A . as the flow rate of pilot diesel is

decreased from 100 t o 13.5 percent of t o t a l energy i n p u t . The
first stage of combustion i n s t r a i g h t d i e s e l o p e r a t i o n consumes
about 70 percent of t h e d i e s e l energy i n p u t (see F i g . 5.17). A
r e d u c t i o n of p i l o t d i e s e l t o 32 percent of total energy input
results in a higher peak f o r r a t e of energy r e l e a s e w i t h no
i n d i c a t i o n of t h e two-staged c o n b u s t i o n . Further reduction of
pilot d i e s e l r e s u l t s i n lower and wider peaks. F i g . 5.18 shows
the f r a c t i o n of f u e l burned a t d i f f e r e n t p i l o t q u a n t i t i e s . The
operation with low p i l o t d i e s e l flow r a t e shows a l a r g e amount
of unburned f u e l . The amount of unburned f u e l d e c r e a s e s as the
flow r a t e of p i l o t d i e s e l i s i n c r e a s e d .

F i g . 5.19 shows t h e r a t e of energy r e l e a s e f o r a h i g h load

operation (brake mean e f f e c t i v e p r e s s u r e of 571 k P a ) . R e d u c t i o n
of p i l o t d i e s e l flow r a t e appears t o i n c r e a s e i g n i t i o n delay and
d e c r e a s e t h e maximum r a t e of energy release. The cumulative
energy release shown i n F i g . 5.20 i n d i c a t e s t h a t a t t h i s l o a d
combustion o c c u r s i n two stages at a l lpilot diesel energy
ratios. The small second-stage combustion i n h i g h l o a d d u a l -
f u e l o p e r a t i o n seems t o be t h e consumption, i n t h e main chamber,
of t h e unburned and/or p a r t i a l l y burned f u e l r e m a i n i n g from the
flame p r o p a g a t i o n i n the f i r s t stage.
 CD
bmep - 279kPa ( 3 3 % full load)

LO
CO
T3.5% diesel
20.6% diesel
.—. cn 32.0% diesel
^ CO
100% diesel

CD
in
(/) CM'
ro
CD
"ai
ai
CN
ai
ro
a> o
m_
cu
>
+->
rd

LP
c6

CD

LP
CD
1 1 "I
•10.0 tdc 10.0 30.0 50.0 70.0

Crank Angle (deg RTQC)

Figure 5.17 - Cumulative Energy Release o f Dual-Fuel Operation at Various Pilot Diesel Flow Rates
Figure 5.18 - Fraction of Fuel Burnt i n Low Load Dual-Fuel Operation
ZZI
 CD

LO
CO

CD

LP
-10.0 tdc 10.0 30.0 50.0 70 0
Crank Angle (deg FITDC)

Figure 5.20 - Cumulative Energy Release of Dual-Fuel Operation at Various Pi

Diesel Flow Rates
 124

5.3.2 E f f e c t of R e s t r i c t i n g I n t a k e A i r

F i g . 5.21 shows the rate of energy r e l e a s e f o r low l o a d

o p e r a t i o n w i t h and w i t h o u t a i r r e s t r i c t i o n . It indicates that
restriction of intake a i r lengthens ignition delay by 1-2
degrees. Table 5.2 i l l u s t r a t e s t h e temperature and p r e s s u r e of
m i x t u r e a t t o p dead c e n t r e w i t h and w i t h o u t r e s t r i c t i o n of a i r .
When the i n t a k e a i r i s r e s t r i c t e d a drop i n the p r e s s u r e a t t o p
dead c e n t r e i s n o t i c e d . A r a t h e r s u p r i s i n g phenomenon is that
there appears from the calculation to be an increase in
temperature. This i s contrary to the assumption that the
temperature would drop as the p r e s s u r e d r o p s , which was t h e
b a s i s used by L e w i s ( l 9 5 3 ) i n e x p l a i n i n g the i n c r e a s e i n i g n i t i o n
d e l a y when the r e s t r i c t i o n of a i r i s imposed. I t seems t h a t the
i n c r e a s e i n i g n i t i o n d e l a y i s not the r e s u l t of the change i n
chemical delay since the-chemical d e l a y would be shortened i f
the temperature i s increased. This would suggest that the
change in ignition delay i s probably more s e n s i t i v e t o t h e
change i n p h y s i c a l d e l a y . F i g . 5.22 shows t h e c u m u l a t i v e energy
release f o r the corresponding operations.

The e f f e c t of a i r r e s t r i c t i o n f o r h i g h load operation is

shown in F i g 5.23 and F i g . 5.24. The o p e r a t i o n a t the brake

mean e f f e c t i v e p r e s s u r e of 571 kPa (67% f u l l load) exhibits a

radical increase i n maximum rate of energy r e l e a s e when t h e

intake a i r i s r e s t r i c t e d . The g a s - a i r m i x t u r e s t r e n g t h f o r the

corresponding operation i s 0.606 i n equivalence ratio. This

v a l u e i s s l i g h t l y above the p r e v i o u s l y mentioned v a l u e f o r the

lower limit of flammability. Thus the sudden i n c r e a s e i n t h e

 CO

co bmep - 279 kPa, a i r r e s t r i c t e d (5 = 0. 400) ,

. . 14.5% d i e s e l g

cm a bmep - 279 kTa, a i r u n r e s t r i c t e d (d = 0.364) ,

QJ
13.5% d i e s e l 9

\
bmep - 143 kPa, a i r r e s t r i c t e d (c5 = 0. 308) ,
1 0

• \ ro _
crj 20.9% d i e s e l g

~J bmep - 143 kPa, a ir u n r e s t r i c t e d (<E 0.279),

19.5% d i e s e l g

QJ ^ .
LO <=>
ro
QJ

>~
CJ)
CU CD
c ;.
UJ <=>
U-
o
QJ ~. .
+-> C D
ro
cc
co
CD

CO
CD

CD
I

-10.0 tdc 10.0 30.0- 50.0 70.0

Crank Angle (deg flTD-C)
Figure 5.21 Effect o f R e s t r i c t i n g Intake A i r on Rate o f Energy Release
LOAD AIR FRACTIONAL D I E S E L * P (kPa) # OF MOLES OF
gas
T
tdc ™ t d c

(kPa) RESTRICTION ENERGY INPUT (%) INTAKE MIXTURE

143 unrestricted 19.5 0. 279 968 4810 0.0631

143 restricted 20.9 0.308 999 4 370 0.0555

279 unrestricted 13.5 0.364 951 4860 0.064 8

279 restricted 14 .5 0.400 1010 44 70 0.0564

4 29 unrestricted 15 .0 0.437 944 5000 0.0673

429 restricted 15 . 3 0.483 1010 4680 0.0586

571 unrestricted 10.2 0.526 955 5210 0.0692

571 restricted 10 . 7 0.606 1010 4 7 8.0 0.06 01

Table 5.2 - Effect o f Intake A i r R e s t r i c t i o n on M i x t u r e Temperature

at Top Dead Center
 CD
bmep - 279 k P a , a i r r e s t r i c t e d (cD = 0 . 4 0 0 ) ,
14.5% d i e s e l g

LO bmep - 279 k P a ,
CO
a i r u n r e s t r i c t e d (cB = 0 . 3 6 4 ) ,
13.5% d i e s e l g

bmep - 1 4 3 k P a , a i r r e s t r i c t e d (<5 = 0 . 3 0 8 ) ,
20.9% d i e s e l g

3 CO bmep - 1 4 3 k P a , a i r u n r e s t r i c t e d (cB 0.279),

19.5% d i e s e l g

CO ^
</> CM
OJ
CO
cc CD
CM'

s-
co
c
LO

CO
>
•r-
+J
—
I CD
3 —'
E
o
LO
CD

CD
CD
CD
1 1
•10.0 tdc 10.0 30.0 50.0 70.0
Crank Angle fdeg flTDC)

Figure 5.22 - E f f e c t of Restricting Intake A i r on Cumulative Energy Release

 Rate of Energy Release (kJ / aeg)
-0.1 -0.G3 0.03 0.1 0.16 0.23 0.3 0.36 0.43 0.5
_] I I I I I I I i I I I I 1 L_

Cu—|
o

CD

O
•J

ID
n
in Gj
• CD nf If tf If
ro "3
i
a

I I
cn
CO

$ 5 5
QJ QJ QJ Q)

Q) t ( QJ P> O J i — i QJ
O P- O P- Ul p-In p-
u-i
a
• 1-1 •
• f< ui
f-(
ZD CP i-( M
o
• t-l
<#> i-i o'P C
•V--- CD
Ch cn
o'P C
Oi cn ^ a
£' R"aP- fDR cn P- P- CD
ft(D cn
a ) cn
tn p- ifl rr n> o cn n-
P" P- P p- 1
:ted

CD Ho<
rb o O

ft ff
-3I
9*
-r-
el
et
il cQ
II

o II
o n
cn o .J>
O
o 00
• •
CO
-—' NJ ' CO

iD —

CD

821
 o

Ln

CD
CO

LO
CM*
01
(O

* CD
cn
429 kPa, a i r unrestricted (c6 =0 . 4 3 7 ) ,
c —
15.0% diesel g

429 k P a , a i r r e s t r i c t e d (« = 0 . 4 8 3 ) ,
> 15.3% diesel g

•r-
+-> —'
CD 571 k P a , a i r u n r e s t r i c t e d (ffi = 0 . 5 2 6 ) ,
UJ 10.2% diesel g

E 571 k P a , a i r r e s t r i c t e d (I> = 0 . 6 0 6 ) ,
3
O LO 10.7% diesel g

CD
CD
in

CD 1—
•10.0 tdc 10.0 30.0 70.0
50.0
Crank Angle (deg ATDC)
Figure 5.24 - Effect of Intake A i r R e s t r i c t i o n on Cumulative Energy Release
 130

maximum r a t e of energy r e l e a s e seems t o be the r e s u l t of fuel-

air ratio on p r o p a g a t i o n of a flame through a premixed g a s - a i r
mixture.

5.3.3 E f f e c t of V a r y i n g I n j e c t i o n Timing

F i g . 5.25 and F i g . 5.26 show t h e e f f e c t of advancing the

i n j e c t i o n t i m i n g a t low l o a d . Advancing the i n j e c t i o n t i m i n g by
10 degree C A . results i n the maximum r a t e of energy release
o c c u r i n g a t t o p dead c e n t e r . For t h i s t i m i n g the second stage
combustion i s more d i s t i n c t . For t h e o p e r a t i o n s w i t h injection
t i m i n g of 12.3 and 17.3 degree C A . BTDC the combustion is
t a k i n g p l a c e o n l y a f t e r t h e p i s t o n has s t a r t e d moving downwards,
thus assisting t h e flow of m i x t u r e from prechamber t o the main
chamber. For the o p e r a t i o n w i t h i n j e c t i o n t i m i n g of 22.3 degree
CA. BTDC, the combustion takes p l a c e when the p i s t o n i s n e a r l y
motionless, and the second stage combustion would be more
distinctive.

The effect of r e t a r d i n g i n j e c t i o n timing at high load i s

shown i n F i g 5.27 and F i g . 5.28. S i g n i f i c a n t reduction i n the
maximum rate of energy r e l e a s e i s observed when t h e t i m i n g i s
retarded. Again the two-staged combustion characteristics
become l e s s d i s t i n c t i v e as the t i m i n g i s r e t a r d e d .
 LTJ
CD

CO bmep - 279 kPa, 20% d i e s e l

. . "3
QJ i n j e c t i o n a t 12.3 BTDC
-a 17.3 BTDC
. CD 22.3 BTDC
\ CO

_l
-10.0
,
tdc
(

10.0
, ,
30.0
r — ,

50.0
,

70.0
r—

Crank Angle (deg ATDCJ

Figure 5.25 - E f f e c t o f Advancing Injection Timing on Rate o f Energy Release
 Cumulative Energy Release (kJ)
-0.5 0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0
_l I I I I I l i i i I I I I ' i i
CD
O

CD
CD

o
QJ

ID
rj .
to co
i , a

1 tr
CL 1 3
70 1
! fD
TI
a i
1 1
i
1

3D ' hj. M

i_J. -J
"'3
a g <£>

?]
rt 13
H-
0 ^
3
0) O
rt <#>

M h- 1
M a
to
• • . ro
CO CO CO cn
(D
h- 1

—1 CD 03 UJ
i-3 t-3 t-3
a

a a a
o n n

ZEI
 C3
LO

CO brep - 571 kPa, 10% d i e s e l

tzn b i n j e c t i o n a t 12.3 BTDC
QJ
"TZJ 10.3 BTDC
8.3 BTDC
CD
-I

CD
QJ ° ? .
LO
ro
QJ

>. CO
OO
P
cn t o
c —. -
LU <=)

CD
QJ —;
•M
ro

CO

CD
OO
CD

-10.0 tdc 10.0 30.0 50.0 70.0

Crank Angle (deg flTDC)

Figure 5.27 - E f f e c t o f Retarding Injection Timing o n Rate o f Energy Relea

 CD

LD
CO

CD

LP

-10.0 tdc 10.0 30.0 50.0 70.0

Crank Angle (deg flTDC)

Figure 5.28 - Effect of Retarding Injection Timing on Cumulative Energy Release

 135

CH.VI C o n c l u s i o n s and Recommendations

6.1 Conclusions

Safe o p e r a t i o n of a prechamber diesel engine with dual-

fuelling with natural gas i s severely limited by maximum
cylinder pressure. Even at half load t h e maximum cylinder
pressure of d u a l - f u e l o p e r a t i o n i s as h i g h as t h a t of s t r a i g h t
d i e s e l operation at f u l l load. The e x c e s s i v e maximum cylinder
pressure i s a s s o c i a t e d w i t h t h e h i g h r a t e of energy r e l e a s e by
combustion which takes p l a c e w i t h i n a n e a r l y homogeneous g a s - a i r
mixture. The maximum c y l i n d e r p r e s s u r e , as w e l l as t h e r a t e of
cylinder pressure rise, can be reduced to a s a f e l e v e l by
r e t a r d i n g the i n j e c t i o n t i m i n g by about 4 t o 6 degrees of crank
angle. The change i n t h e r m a l e f f i c i e n c y due t o t h e r e t a r d a t i o n
i s s m a l l ( l e s s than 0.5 p e r c e n t with 4 degree retardation at
67 p e r c e n t of f u l l load). R e s t r i c t i n g t h e i n t a k e a i r can reduce
also t h e maximum c y l i n d e r p r e s s u r e , but t h i s r e s u l t s i n h i g h e r
maximum r a t e of c y l i n d e r p r e s s u r e rise.

S t a b l e d u a l - f u e l o p e r a t i o n r e q u i r e s s u f f i c i e n t f l o w r a t e of

pilot diesel fuel; insufficient amount, of pilot diesel fuel

results i n e r r a t i c operation with m i s f i r e d c y c l e s . The minimum

p i l o t d i e s e l f u e l r e q u i r e d i n o r d e r t o ensure a s t a b l e o p e r a t i o n

is typically 8 to 15 percent of the t o t a l energy input,

depending on engine l o a d .
 136

Dual-fuel operation at p a r t l o a d showed g e n e r a l l y higher

f u e l consumption than t h a t of s t r a i g h t diesel operation. The
apparent energy release a n a l y s i s revealed t h a t the h i g h e r fuel
consumption r a t e i s m a i n l y due t o gas s u r v i v i n g unburned t h r o u g h
the. combustion chamber. The main cause of t h i s poor combustion
i s weak g a s - a i r m i x t u r e s t r e n g t h . Increase i n p i l o t d i e s e l flow
rates reduces the amount of unburned gas and thus improves the
f u e l consumption r a t e . The dependence of t o t a l f u e l consumption
r a t e on p i l o t d i e s e l flow rate is less with higher gas-air
mixture concentration. With r e s t r i c t i o n on t u r b o c h a r g e r i n l e t
pressure ( t o p r e v e n t surge) less than about 10 percent a i r
restriction was possible in the t e s t s c o n d u c t e d ; t h i s showed
some improvement (perhaps one p e r c e n t ) on f u e l consumption r a t e .
Advancing i n j e c t i o n t i m i n g showed no s i g n i f i c a n t e f f e c t on fuel
consumption r a t e .

The fuel consumption r a t e d u r i n g d u a l - f u e l o p e r a t i o n near

full load approached that of straight diesel operation.
Extrapolated beyond full load diesel operation, the fuel
consumption r a t e of d u a l - f u e l o p e r a t i o n would become lower than
t h a t of s t r a i g h t d i e s e l o p e r a t i o n .

The combustion c h a r a c t e r i s t i c s of s t r a i g h t d i e s e l and dual-

fuel operation differ in that in the former the combustion

c o n s i s t s m a i n l y of a u t o - i g n i t i o n of d i e s e l f u e l , whereas i n the

l a t t e r the combustion i s c a r r i e d out by the p r o p a g a t i o n of flame

fronts. T h i s d i s t i n c t i o n was c l e a r l y e x h i b i t e d i n the a n a l y s i s

of apparent energy release, where straight diesel operation

showed a two-staged combustion and dual-fuel operation a

 137

s h o r t e r , s i n g l e - s t a g e d combustion.
 138

6.2 Recommendations

• Without d i f f i c u l t modifications the c u r r e n t engine can be

c o n v e r t e d i n t o a d i r e c t - i n j e c t i o n engine. F u r t h e r t e s t s on t h i s
converted engine would lead to comparative study between
prechamber and d i r e c t - i n j e c t i o n engines.

• Further studies at d i f f e r e n t engine speeds a r e r e q u i r e d .

• F u r t h e r study of t h e e f f e c t of r e s t r i c t i n g a i r i n t a k e without
the turbocharger may show some improvement i n f u e l consumption
at low l o a d s .

• Exhaust gas a n a l y s i s would p r o v i d e some important information

such as t h e lower l i m i t of f l a m m a b i l i t y .
 139

BIBLIOGRAPHY

1. ANNAND, W.J.D., Heat T r a n s f e r i n the C y l i n d e r s of

R e c i p r o c a t i n g I n t e r n a l Combustion E n g i n e s " , Proc. Inst. of
Mech. Eng., V o l . 177,No. 36, 1963.

2. AUSTEN,A.E.W.,LYN,W.T., " R e l a t i o n between F u e l I n j e c t i o n

and Heat R e l e a s e i n a D i r e c t - I n j e c t i o n Engine and the Nature of
Combustion P r o c e s s e s " , P r o c . Auto. D i v . I n s t , of Mech. Eng.
No.1,1960-61
3. BENSON, R.S.,WHITEHOUSE,N.D., "Internal Combustion
E n g i n e s " , V o l 1&2, Pergamon P r e s s 1979
4. BOLT, J.A., HENEIN, N.A., "The E f f e c t of Some Engine
V a r i a b l e s on I g n i t i o n Delay and Other Combustion Phenomena i n a
D i e s e l Engine", Proc. I n s t , of Mech. Eng., V o l . 184, P t . 3 J ,
p p U O , 1 969/70.

5. BOYER, R.L., S t a t u s of D u a l - F u e l Engine Development, Paper

to S.A.E. Annual M e e t i n g , D e t r o i t , January I 0 t h - 1 4 t h , 1949.
6. BOYER, R.L., Crooks, W.R., "The Modern Gas E n g i n e s " , ASME,
Paper No. 51-OGP-4, 1951.
7. CAMPBELL,A.S., "Thermodynamic A n a y s i s of Combustion
E n g i n e s " , John W i l e y & Sons, 1979
8. CHIANG, C.W., MYERS, P.S., VYEHARA, O.E., "Physical and
Chemical I g n i t i o n D e l a y " , SAE Trans. V o l . 68, 1960.
9. ELLIOTT, M.A., DAVIS, R.F., " D u a l - F u e l Combustion i n D i e s e l
E n g i n e s " , Ind. and Eng. Chem., V o l . 43, No. 12, 1951.
10. FELT,A.E.,STEELE,A.E.,Jr., Combustion C o n t r o l i n Dual-Fuel
E n g i n e s , SAE T r a n s . , V o l . 7 0 , p644-653, 1962
11. GEE,D.E. and KARIM,G.A., "Heat Release i n a Compression-
I g n i t i o n E n g i n e " , E n g i n e e r , Lond.~, 1 966 (September ) .
12. HOPKINS, V.H.F., TALBUTT, R.J., WHITEHOUSE, N.D., "The
Development of D u a l - F u e l Engines f o r Powers above 1,5000 bhp",
English Elec. J . V o l . 1 8 , No. 3, May-June, 1963.

13. JONES, J . , "The P o s i t i o n and Development of the Gas

Engine", P r o c . I n s t . of Mech. Eng., V o l . 151, pp32,l944.

14. KAMEL, M., WATSON, N., "Heat T r a n s f e r i n The Indirect

I n j e c t i o n D i e s e l E n g i n e " , SAE Paper No. 790826, 1979.

15. KARIM, G.A., KLAT, S.R., MOORE, N.P.W.," "Knock i n D u a l -

Fuel Engines", Proc. Inst. of Mech. Eng., V o l . 1 8 1 , P t . 1,
No. 20, 1966/67.
 140

16. KARIM, G.A.,KHAN, M.O., "A G e n e r a l i z e d Computer Programm

f o r t h e D e t e r m i n a t i o n of Rates of Heat Release and Heat T r a n s f e r
i n R e c i p r o c a t i n g E n g i n e s " , I m p e r i a l C o l l e g e , London, Dec. 1967.

17. KARIM, G.A., The I g n i t i o n of a Premixed F u e l and A i r

Charge by P i l o t ^ F u e l Spray I n j e c t i o n w i t h Reference t o D u a l - F u e l
Combustion, SAE T r a n s . V o l . 7 7 , 1968

18. KARIM, G.A.,KHAN, M.O., "Examination of e f f e c t i v e r a t e s of

Combustion Heat Release i n a D u a l - F u e l Engine", J o u r n . Mech.
Eng. Sc., V o l . 10, pp13. 1968

19. KARIM, G.A., A Review of Combustion P r o c e s s e s i n the Dual

F u e l E n g i n e - t h e Gas D i e s e l Engine, Prog. Energy Combust. S c i . ,
v o l . 6 , P277-285, 1980
20. KARIM, G.A., BURN, K.S., "The Combustion of Gaseous F u e l s
i n a Dual F u e l Engine of t h e Compression I g n i t i o n Type w i t h
P a r t i c u l a r Reference t o C o l d I n t a k e Temperature C o n d i t i o n s " ,
SAE, Paper No. 800263, 1980.

21. KARIM, G.A., Methane and D i e s e l E n g i n e s , Proc. of

"Methane-Fuel f o r t h e F u t u r e . " , p113-129, 1982
22. KARIM, G.A., "Some C o n s i d e r a t i o n s of t h e Use of N a t u r a l
Gas i n D i e s e l E n g i n e s " , Symposium on Nonpetroleum V e h i c u l a r
F u e l s I I I , A r l i n g t o n , V i r g i n i a , O c t . 12-14, 1982.
23. KARIM, G.A.,WIERZBA, I . , "Comparative S t u d i e s of Methane
and Propane as F u e l s f o r S p a r t I g n i t i o n and Compression I g n i t i o n
E n g i n e s " , Paper t o S.A.E. West Coast International Meeting
Vancouver, B r i t i s h Columbia, Aug. 8-11, 1983.
24. KELGARD, E., " O p e r a t i o n of D u a l - F u e l Engines i n P i p e l i n e
S e r v i c e " , ASME, Paper No. 61-OGP-1, 1961.
25. KRIEGER,R.B., BORMAN,B.L., "A g e n e r a l i z e d Computer
Programme f o r the D e t e r m i n a t i o n of Rates of Heat Release and
Heat T r a n s f e r i n R e c i p r o c a t i n g E n g i n e s " , Imperial
C o l l e g e , L o n d o n , Dec. 1967.

26. LEWIS,J.D., A Study of Combustion P r o c e s s e s i n the D u a l -

f u e l Engine by Exhaust Gas A n a y s i s Methods, Ph.D. Thesis,
London U n i v e r s i t y , 1954
27. LYN, W.T., " C a l c u l a t i o n of the E f f e c t of Rate of Heat
Release on t h e Shape of C y l i n d e r - P r e s s u r e Diagram and C y c l e
Ef f i c i e n c y " , P r o c . I n s t . of Mech. Eng., No. 1,pp34, 1960/61.
28. LYN,W.T., "Study of B u r n i n g Rate and Nature of Combustion
i n D i e s e l E n g i n e s " , N i n t h I n t e r n a t i o n a l Symposium on Combustion,
Academic P r e s s , New York, N.Y., p p l 0 6 9 - l 0 8 2 , 1962.

29. LYN,W.T., SAMAGA,B.S., BOWDEN,C.M., "Rate of Heat R e l e a s e

i n High-Speed I n d i r e c t - I n j e c t i o n D i e s e l E n g i n e s " , P r o c . Inst.
Of Mech.Eng., V o l . 184, P t . 3 J , pp122, 1969/70.
 141

30. MITCHELL, R.W.S., and WHITEHOUSE,N.D., The Development and

Performance of a Range of Dual F u e l E n g i n e s , Canadian M i n i n g and
M e t a l l u r g i c a l B u l l e t i n , 1955

31. Moore, N.P.W., and MITCHELL,R.W.S., Combustion i n Dual

F u e l E n g i n e s , J o i n t Conf. on Combustion, ASME/Inst. of Mech.
Eng. pp300 1955.

32. MURAYAMA,T,MIYAMOTO,N,and FUKAZAWA,S, An E x p e r i m e n t a l

Study on t h e Performance o f a M u l t i f u e l Engine, Bull. JSME
Vol.14 N.67, p76-83, 1971

33. OBERT,E.F., I n t e r n a l Combustion Engines and A i r P o l l u t i o n ,

3rd ED., Harper & Row, 1973

34. O'NEAL,. G.B., "The D i e s e l - G a s D u a l - F u e l E n g i n e " , Symposium

on Nonpetroleum V e h i c u l a r F u e l s I I I , A r l i n g t o n , Virginia,
Oct. 12-14, 1982.

35. ROUGHTON, J.H., " D u a l - F u e l E n g i n e s " , S u r v e y o r , V o l . 135,

No. 4064,pp34-8, May, 1970.

36. SIMONSON, J.R., "Some Combustion Problems of the D u a l - F u e l

Engine", E n g i n e e r i n g , v o l . 178, p.363, 1954.
37. SIMONSON, J.R., "An A n a l y s i s of D u a l - F u e l Combustion
Proceses i n a Compression I g n i t i o n E n g i n e " , Ph.D T h e s i s , London
U n i v e r s i t y , 1955.
38. STEVEN. G., " O p e r a t i n g Problems of the D u a l - F u e l E n g i n e " ,
Power E n g i n e r r i n g , March, 1953.
39. TAYLOR, C.F., "The I n t e r n a l Combustion Engine i n Theory
and P r a c t i c e " , V o l . 1 & 2, M.I.T. P r e s s , 1982.

40. VAN WYLEN, G.J., SONTAG, R.E., Fundamental of C l a s s i c a l

Thermodynamics", 2nd ED., John W i l e y & Sons, 1978.
41. WATSON, N.,KAMEL, M., "Thermodynamic E f f e c i e n c e e v a l u a t i o n
of an I n d i r e c t I n j e c t i o n D i e s e l E n g i n e " , SAE Paper No. 790039,
Feb-Mar 1979.

42. WHITEHOUSE, N.D., STOTTER, A., GOUDIE, G.D., PRENTICE,

B.W., "Method of P r e d i c t i n g Some A s p e c t s of Performance of a
D i e s e l Engine U s i n g a D i g i t a l Computer", P r o c . I n s t . of Mech.
Eng., V o l . 176, No, 9, 1962.
43. WHITEHOUSE, N.D., WAY, R., Rate of Heat R e l e a s e i n D i e s e l
Engines and I t ' s C o r r e l a t i o n w i t h F u e l I n j e c t i o n Data", Proc.
I n s t . Mech. Eng., V o l . 184, P t . 3 J , p p l 7 , 1969/70.
 142

APPENDIX A - CALIBRATION CURVES

Load Sensor

o
o"

o
o'
o

o
o'
00

o
o'

T3

O
•J
O

Voltage (mV)

* slope - 22.2 N/mV

 143

Air Flow Element

 144

Gas Flow Element

Differential Pressure (kPa)

145
 146

APPENDIX B - COMPUTATION OF INDICATED MEAN EFFECTIVE PRESSURE

Definition

An indicated mean effective pressure, imep, i s defined

as that theoretical constant pressure which c a n be imagined

exerted during each power stroke of the engine to produce

work equal to the i n d i c a t e d work:

P dV = P. „ AV
ind
J V
l

where P - cylinder pressure

P. , - i n d i c a t e d mean effective
ind
- pressure, imep

V - cylinder volume

V^ - V at the beginning of
the cycle

- V a t t h e end o f t h e cycle

bdc tdc

Indicated mean effective pressure, in this project, was

computed according to the following:

(P. + P.,,)
1 + 1
( V . „ - V.)
2 i+1 r

^ ind
( V
bdc V
tdc )

where P^, are cylinder pressure and

volume at i ^ t 1
stage, and each

stage corresponds to a degree of

crank angle throughout a cycle.

147
 148

APPENDIX C - COMPUTER PROGRAM FOR CYLINDER PRESSURE DATA

ACQUISITION
 149

Listing of APP.PR0G2 a t 22:10:05 on APR 11, 1984 f o r CCid=AFPH

1 C
2 C Acquires data from D i e s e l Engine
3 C
4 EXTERNAL QTQIO
5 EXTERNAL GETADR
6 EXTERNAL ASNLUN
7 C
e9 INTEGER
INTEGER
L I S T C 3 0 0 2 ) , IDAT(3002), 1PARM(6)
YES, NO, ANS, IPARR(5), IBDC(3)
ISTAT(2)

10 REAL LOAD, PMEAN(721), VOLUME*5)

1 1 C
12 YES = 1
1 3 NO = 0
1 4 SCALE = -32768.0
1 5 NPOINT = 3001
1 6 NPCYC = 720
1 7 NPCYCE = 750
18 NPCYCS = 700
19 NPCY2 = 360
20 STROKE = 6.0
21 ARM = 3.0
22 ROD = 9.595
23 VCLEAR = 6.444
24 PTAREA = 3.1416* ( 4 . 7 5 / 2.0)** 2
25 NPC1 = 721
26 FNPC1 - FLOAT(NPC1)
27 NPCY3 = NPCY2 + 50
28 RDPDEG = 0.0174533
29 PTHR = 400.0
30 PMINRF = 14.7
31 STHR = 0.6
32 C
33 DO 5 1 = 1 ,3002
34 IDAT(I) = 0
35 5 CONTINUE
36 C
37 DO 6 1=1,NPC1
38 PMEAN(I) = 0.0
39 6 CONTINUE
40 CALL ASNLUN(3, 'NI', 0)
4 1 CALL ASSIGN*1, 'PAR2.DAT')
42 C
43 READ(1,100) CLOCK
44 DWELL = 1. / CLOCK
45 HERTZ = 1 . / XRATE(DWELL, I RATE, IPRSET, 1)
46 CALL CLOCKB(I RATE, IPRSET, 1, IND, 1)
47 DELT = 1. / HERTZ
48 WRITE(5,200) IND, HERTZ
49 C
50 READ(1,101) NCHAN
51 READ(1,102) NDPCH
52 C
53
54
c NDTOT = NPOINT
55 ISTRT = 1
56 I LAST = NDTOT + 1
57
58
c read in port address
READ(1,103) LIST(1)
 150

Listing of APP.PROG2 a t 22:10:05 on APR 11 , 1 984 f o r CCid=AFPH

59 c r e a d i n scan i n s t r u c t i o n s
60 DO 10 I=2,NCHAN+1
61 READ(1,103) L1STU)
62 10 CONTINUE
63 c f i l l the r e s t of scan l i s t by r e p e a t i n g
64 DO 11 I=NCHAN+2,ILAST
65 L I S T ( I ) " = LIST(I-NCHAN)
66 11 CONTINUE
67 c r e s e t s e r i e s 500 BUS
6B IPARMC2) = 2
69 CALL GETADR(IPARM(1), IDAT)
70 CALL WTQ10("1002,3,10,1,1 STAT,IPARM,IDS)
71 WRITE(5,201)
72 WRITE(5,202) I S T A T ( I ) , I S T A T ( 2 ) , IDS
73 c
74 IRSA = 1
75 c w r i t e scan l i s t to RAM, read back and c h e c k
76 IPARM(2) = (ILAST-ISTRT+1) * 2
77 IPARM(3) = IRSA
78 CALL GETADR(IPARM(1),LI ST(1 START))
79 CALL WTQIO("400,3, 10, 1 ,I STAT,IPARM,IDS)
80 c read back
81 CALL • GETADR(IPARM(1), I DAT(ISTRT))
82 CALL WTQIO("1000,3,10,1,1 STAT,IPARM,IDS)
83 c p r i n t any d i s c r e p a n c i e s
84 I ERR = 0
85 DO 20 1 = 1STRT,I LAST
86 IF (IDAT(I) .EQ. L I S T ( I ) ) GO TO 20
87 I ERR = I ERR + 1
88 WRITE(5,203) LI ST(I ),I DAT(I)
89 20 CONTINUE
90 WRITE(5,204 ) I ERR
91 WRITE(5,202) I S T A T ( l ) , I S T A T ( 2 ) , IDS
92 IF (I ERR .GT. 0) GO TO 999
93 c a c q u i r e data
94 IWCT = NDTOT + 1
95 CALL I DATE(ID 1 , ID2, ID3)
96 WR1TE(5,205) ID1, ID2, ID3
97 c
98 c calibration of P r e s s u r e Measurement
99 c
100 PCPPSI = 0.830
101 CHMUPV = 1000.0
1 02 PCPMU = 1.415
103 GAIN =1.0
104 C1P = CHMUPV * PCPMU / PCPPSI / SCALE * GAIN
1 05 c
106 DO 600 1600=1,100
107 WRITE(5,270)
108 READ(5,170) SPEED
109 FREQRQ = SPEED / 60.0 / 2.0 * FLOAD(NPCYC) * 2
11 0 DWELL = 1. / FREQRQ
11 1 HERTZ = 1. / XRATE(DWELL, I RATE, IPRSET, 1)
1 12 CALL CLOCKBCIRATE, IPRSET, 1, IND, 1)
11 3 DELT = 1. / HERTZ * 2
1 14 DTPDEG = DELT
1 15 WRITE(5,200) IND, HERTZ
1 16 NCYCLE =10
 151

Listing of APP.PROG2 at 22:10:05 on APR 11, 1984 f o r CCid=AFPH

1 7 WRITE(5,271) NCYCLE
1 8 READ(5,171) NCYCLE
1 9 IPMAX = 0
20 FNCYC = FLOAT(NCYCLE)
21 NSET = 1
22 DO 610 1610=1,NPC1
23 PMEAN(1610) = 0.0
24 610 CONTINUE
25 PMAX0 = 0.0
26 DPDTM0 = 0.0
27 RPMIN0 = 0.0
28 RIMEP0 = 0.0
29 SUMRP = 0.0
30 SUM2RP = 0.0
31 SUMIM = 0.0
32 SUM2IM = 0.0
33 SUMPX = 0.0
34 SUM2PX = 0.0
35 SUMPN = 0.0
36 SUM2PN = 0.0
37 SUMDN = 0.0
38 SUM2DN = 0.0
39 WRITE(5,269)
40 READ(5,170) PINTAK
41 PINTAK >= PINTAK + 14.7
42 WRITE(5,210)
43 READ(5,110) CR
44 DO 650 M650=1,500
45 IF (NSET .GT. NCYCLE) GO TO 651
46 C
47 CALL GETADR(I PARK(1 ) , I DAT)
46 IPARM(2) = IWCT *2
49 IPARM(3) = IRSA
50 CALL WTQIOC3001 ,3,1 0, 1 ,I STAT,IPARM,IDS)
51 C
52 MB = 3
53 MF =751
54 DO 680 M680=1 ,2
55 DO 661 M=MB,MF,2
56 S = IDAT(M)
57 S = ABS(S/SCALE)
58 IF (S .LT. STHR) GO TO 681
59 P = IDAT(M+359)
60 P = P * C1P
61 IF (P .LT. PTHR) GO TO 68 2
62 IBDC(1) = M
63 GO TO 683
64 682 CONTINUE
65 MB = M + 680
66 MF = MB + 80
67 GO TO 680
66 681 CONTINUE
69 680 CONTINUE
70 GO TO 650
71 663 CONTINUE
72 INDXBD = 2
73 MB = IBDC(1) + 680
74 MF = MB + 80
 152

Listing of APP.PR0G2 a t 22: 10:05 on APR 1 1, 1984 f o r CCi d=AFPH

175 DO 685 M685=1,2

176 DO 686 M=MB,MF,2
177 S = 1DAT(M)
178 S = ABS(S/SCALE)
179 IF (S .LT. STHR) GOTO 686
180 IBDC(INDXBD) = M
181 MB = M + 680
182 MF = MB + 80
183 INDXBD = INDXBD + 1
184 GO TO 685
185 686 CONTINUE
186 GO TO 650
187 685 CONTINUE
188 IDIFFA = IBDC(2) - I B D C ( 1 )
189 IDIFFB = IBDC(3) - IBDC(2)
190 MB = IBDC(2) + 681
191 MF = MF + 80
192 PMIN = 0.0
193 621 CONTINUE
194 IDIFF = IDIFFA + IDIFFB
195 RPMIND = 60.0 / (FLOAT(IDIFF) * DELT) * 4.0
196 MM = IBDC(3) - 81
197 DO 625 M= 1 ,40
198 PM = I DAT(MM)
199 PMIN = PMIN + PM
200 MM = MM + 2
201 625 CONTINUE
202 PMIN = PMIN / 40.0 *C1P
203 626 CONTINUE
204 IBDC1P = IBDC(1) - 1
205 IBDC3P = IBDC(3) - 1
206 PMAX = 0.0
207 DPDTMX=0.0
208 RIMEP =0.0
209 THETA =-180.0
210 DTHETA = 720. / FLOAT(IDIFF) * 2.0
211 DO 632 J1=IBDC1P,1BDC3P,2
212 PJ1P1 = IDAT(J1+2)
213 PJ1 IDAT(J1)
214 PJ1P1 - PJ1P1 * C1P - PMIN + PINTAK
215 PJ1 = PJ1 * C 1 P - PMIN + PINTAK
216 DPDT = ABS((PJ1P1-PJ1) / DELT) * DTPDEG
217 IF (DPDT .GT. DPDTMX) DPDTMX = DPDT
218 RAD1 = THETA * RDPDEG
219 RAD2 = (THETA + DTHETA) * RDPDEG
220 X1 = -ARM * COS(RADI)
221 1 SQRT(ROD*ROD - (ARM * SIN(RADI)) ** 2)
222 X2 = -ARM * COS(RAD2)
223 1 ~ SQRT(ROD*ROD - (ARM * SIN(RAD2)) ** 2)
224 PAVER = (PJ1 + PJ1P1) / 2.0
225 RIMEP = RIMEP + PAVER / STROKE * (X2 - X1)
226 I F (PJ1 .GT. PMAX) PMAX = PJ1
227 THETA = THETA + DTHETA
228 632 CONTINUE
229 WRITE(5,914) IBDC1P, IBDC3P
230 914 FORMAT*' C y c l e l i e s between ',15,' < > ',15,' d e g r e e s ')
231 WRITE(5,272) NSET, ID1FF, RPMIND
232 WRITE(5,273)
 153

Listing of APP.PROG2 a t 22:10:05 on APR 11, 1984 f o r CCid=AFPH

233 WRITE(5,274) PMAX, PINTAK, RIMEP, DPDTMX

234 C
235 WRITE(5,279)
236 READ(5,120) ANS
237 IF (ANS .EQ. NO) GO TO 650
2 38 FI BDC 1 «= FLOAT(IBDCIP)
239 DO 637 JI=1,721
240 F J J 1 - F L O A T ( J I - l ) / DTHETA * 2.0 + FIBDC1
241 INDX - ( F J J 1 / 2.0)
242 INDX - INDX * 2
243 RINDX = (FJJ1 - FLOAT(INDX)) / 2.0
244 PINDX = IDAT(INDX)
245 PINDX1 = IDAT(INDX+2)
246 PINDX = PINDX *C1P - PMIN + PINTAK
247 PINDX1 = PINDX1 * C1P - PMIN + P.1NTAK
248 P = PINDX * RINDX * (PINDX1 - PINDX)
249 PMEAN(J1) = PMEAN(J1) + P / FNCYC
250 637 CONTINUE
251 PMAX0 = PMAX0 + PMAX
252 RIMEP0 = RIMEP0 + RIMEP
253 DPDTM0 = DPDTM0 + DPDTMX
2 54 RPMIN0 = RPMIN0 + RPMIND
2 55 SUMRP = SUMRP + RPMIND
256 SUM2RP « SUM2RP + RPMIND*RPMIND
257 SUMIM « SUMIM + RIMEP
258 SUM2IM = SUM2IM + RIMEP*RIMEP
259 SUMPX n SUMPX • PMAX
260 SUM2PX - SUM2PX + PMAX*PMAX
261 SUMPN • SUMPN + PMIN
262 SUM2PN = SUM2PN + PMIN*PMIN
263 SUMDP = SUMDP + DPDTMX
264 SUM2DP = SUM2DP + DPDTMX*DPDTMX
265 NSET = NSET + 1
266 650 CONTINUE
267 WRITE(5,299)
268 GO TO 652
269 651 CONTINUE
270 PMAX0 = PMAX0 / FNCYC
271 DPDTM0 ' DPDTMO / FNCYC
272 RPMIN0 « RPMINO / FNCYC
273 RIMEPO « RIMEPO / FNCYC
274 FNCYC1 = FNCYC - 1
275 IF (FNCYC1 .EQ. 0) FNCYC1=1
276 SDIMEP = SQRT((SUM2IM-SUMIM*SUMIM/FNCYC)/FNCYC1)
277 SDPMAX •= SQRT((SUM2PX-SUMPX*SUMPX/FNCYC)/FNCYC1)
278 SDPMIN «= SQRT( (SUM2PN-SUMPX*SUMPN/FNCYC )/FNCYC 1 )
279 SDDPDT = SQRT((SUM2DP-SUMDP*SUMDP/FNCYC)/FNCYC1)
280 SDRPM = SQRT((SUM2RP-SUMRP*SUMRP/FNCYC)/FNCYC1)
281 WR1TE(5,275)
282 WRITE(5,276) NCYCLE, SPEED, HERTZ
283 WRITE(5,277)
284 WRITE(5,27B) RPMINO, PMAXO, PINTAK, RIMEPO, DPDTMO
2B5 WRITE(5,286) SDRPM, SDPMAX, SDPMIN, SDIMEP, SDDPDT
286 C
287 652 CONTINUE
288 WRITE(5,290)
289 READ(5,120) ANS
290 I F (ANS .EQ. NO) GO TO 998
 154

Listing of APP.PROG2 a t 22:10:05 on APR 1 1, 1984 f o r CCi d = AFPH

291 600 CONTINUE

292 C
293 C
294 998 CONTINUE
295 WRITE(5,29l)
296 READ(5,120) ANS
297 IF (ANS •EQ. NO) GO TO 999
298 CALL ASSIGN(2, 'P.DAT')
299 WRITE(5,292)
300 READ(5,192) NDEG
301 IPVDGM = NO
302 INTERD = NDEG / 180
303 IF (IPVDGM .EQ. YES) INTERD = INTERD * 2
304 NDAT = NDEG / INTERD + 1
305 NDEG2 = NDEG / 2
306 ITDC = 181
307 IPBEG = ITDC - NDEG2
3 08 IPEND = ITDC + NDEG2
309 IF '(NEG .EQ. 720) IPBEG = 1
310 IF (NEG .EQ. 720) IPEND = 721
311 WRITE(2,700)
312 PMAX = 0.0
313 NSKIP = 5 * INTERD
314 DO 660 1660=IPBEG,IPEND,NSKIP
315 JB = 1660
316 J F - 1660 + NSKIP -1
317 JL = 1
318 DO 661 J=JB,JF,INTERD
3 19 P = PMEAN(J)
320 IPARR(JL) = P
321 JL = JL + 1
322 661 CONTINUE
323 JLM1 = J L -1
324 WRITE(2,70l) (IPARR(LL), LL=1,JLM1)
325 660 CONTINUE
326 MINX = IPBEG - 181
327 MAXX = IPEND - 181
328 WRITE(2,702)
329 IF (IPVDGM .EQ. YES)- GO TO 670
330 WRITE(2,703) NDAT, MINX, INTERD
331 WRITE(2,704) MINX, MAXX
332 WRITE(2,705)
333 WRITE(2,706) RPMIN0
334 WRITE(2,707)
335 GO TO 674
336 670 CONTINUE
337 WRITE(2,710)
338 THETA = MINX
339 FINTER = FLOAT(INTERD)
34 0 THETA = THETA - FINTER
•34 1 DO 662 I 662 = IPBEG,IPEND,NSKIP
342 JB = 1662
343 J F = 1662 + NSKIP -1
344 JL = 1
345 DO 663 J=JB,JF,INTERD
34 6 THETA = THETA + FINTER
34 7 RAD = THETA * RDPDEG
348 X «= ROD + ARM - ARM * COS(RAD)
 155

Listing of APP.PROG2 a t 22: 10:05 on APR 1 1 , 1964 f o r C C i d = AFPH

349 1 - SQRT(ROD * ROD - (ARM * SIN(RAD)) ** 2)

3 50 VOLUME(JL) = X * PTAREA + VCLEAR
351 J L = JL + 1
352 663 CONTINUE
353 JLM1 = J L - 1
354 WRITE(2,711) (VOLUME(LL), LL=1,JLM1)
355 662 CONTINUE
356 WRITE(2,702)
357 WRITE(2,714)
356 WRITE(2,715)
359 WRITE(2,706) RPMINO
360 WRITE(2,717)
361 674 CONTINUE
362 WRITE(2,708)
363 WRITE(2,709)
364 999 CONTINUE
365 STOP
366 C
367 C
366 100 FORMAT(1X,F16.5)
369 101 FORMAT(1X,I 2)
370 102 FORMAT(1X,I 4)
371 103 FORMAT(05)
372 110 FORMAT(A5)
373 120 FORMAT(11)
374 150 FORMAT(F12.4)
375 170 FORMAT(F7.2)
376 17 1 FORMAT(I 2)
377 180 FORMAT(F5.1,1X.F5.3,IX,F5.1)
378 192 FORMAT(13,1X,I 2)
379 195 FORMAT(II)
380 C
381 C
382 200 FORMAT(IX,' IND CODE = ', 13,', F r e q u e n c y = ' , F 8 . 1 , ' Hz')
383 201 FORMAT(1X,' SERIES 500 BUS RESET!!!',/)
384 202 FORMAT(1X,' DRIVER COMPLETION CODE =',06,' (OCTAL)',/
385 1, 1X,' LAST RESPONSE =',06,' (OCTAL)',/
386 2, 1X,' DIRECTIVE STATUS =',06,' (OCTAL)',/)
387 203 FORMAT(IX,'XXXX RAM ERRORXXXXXXXXXXXXXXXXXXXXXXXXXX',/
388 1, 1X,' OUTPUT = ',05,' ; READ BACK = ',05,/)
389 204 FORMAT(1X,' WRITE TO RAM AND READBACK COMPLETE*,I 3,' ERRORS',
390 205 FORMAT(//,IX,' d a t e : ' , I 3,1X,I 2,IX,I 2,//)
391 210 FORMAT(IX,'?? To s t a r t s c a n i n g , e n t e r RETURN!',$)
392 225 FORMAT(1X,I5,2X,E14.6)
393 269 FORMAT(IX,' >> > Enter Intake P r e s s u r e i n ( p s i ) : ' , $ )
394 270 FORMAT(IX," >> > Enter Engine Speed i n RPM:',$)
395 271 FORMAT(1X,' I d e a l * o f C y c l e s i s ',12,', E n t e r d e s i r e d * 12 ',
396 272 FORMAT(10X,' # of d a t a p o i n t s i n ',13,'th c y c l e i s ',14,/,
397 1 10X,' I n d i c a t e d E n g i n e Speed i s ',F6.1,' rpm')
398 273 FORMAT(IX,' P max ( p s i ) P i n t a k e IMEP dpdt max ( p s i / d e g
399 274 FORMAT(1X,3(2X,FB.3),4X,F11.2)
400 275 F0RMAT(15X,' # of C y c l e s E n g i n e Speed Data a q u s i t F r e q ' , / ,
401 1 15X,' (cycles) (rpm) (Hz) ')
402 276 FORMAT(15X,BX,I 2,3X,5X,F6.1,4X,7X,F8.1)
403 277 FORMAT(IX,' Mean: I n d i c a t e d Speed P max P i n t a k e IMEP
404 1 ' dPdt max ')
405 278 FORMAT(IX,1IX,F8.3,5X,5(F11.2))
406 279 FORMAT(IX,' Do you want t o s e l e c t t h i s c y c l e ? ( 1 / 0 ) : ' , $ )
 156

Listing o f APP.PP0G2 a t 2 2 : 1 0 : 05 on APR 11, 1984 f o r C C i d = AFPH

407 2 8 0 F O R M A T ( 1 X , ' >> > E n t e r M.U. p e r V , p C p e r M.U., G a i n : ' , ? )

408 ~ 8 u F O R M A T * I X , ' S t a n d . D e v L ' , F 8 . 3 , 5X , 5 ( F 1 1 . 3 ) )
409 2 5 0 F O R M A T ( / / , ? ? ? ? ? ? To R e r u n e n t e r l o r 0 ' , S )
1

410 2 9 : FORMAT(1X,'To save d a t a f o r p l o t , E n t e r 1 o r 0 : ' , S )

411 ? 9 2 FORMAT!1X,'Enter Crank A n g l e Range 13:',$)
412 7 9 5 F O R M A T * I X . ' D o y o u w a n t a P-V d i a g r a m ? Enter l or 0 : ' , S )
413 299 FORMAT*' <======> C y c l e N O T f o u n d . <======> ')
414 7 00 FORMAT(1X,'EN PRES f)
4 15 710 F O R M A T * I X , ' E N VOL &')
4 16 701 FORMAT*IX,5(15,IX),'I')
417 7 11 FORMAT*1X,5(F5.1,1X),'')
418 7 02 FORMAT(IX,';')
4 19 "i*3 F O R M A T ( I X , ' E N A N G L SHOR ' , 1 3 , I X , 1 4 , I X , 1 2 )
420 7;.4 F O R M A T ( 1X , ' GR A N G L P R E S ; YR - 2 0 0 1 6 0 0 ; X R ' , I 4 , 1X , I 3 , ' ; £ ' )
421 7U F O R M A T ( I X , ' GR V O L P R E S ; YR - 2 0 0 1 6 0 0 ; X R 5 1 1 5 ; i ' )
422 705 FORMAT ( l X , ' T I ''Cylinder pressure vs C r a n k A n g l e ' ';(,')
423 7 •, 5 FORMAT ( l X , ' T I ' ' C y l i n d e r p r e s s u r e vs Volume'';i')
424 706 FORMAT*IX,'DA ' ' ' , F 6 . l , ' rpm'';&')
425 707 FORMAT ( I X , ' X T I T ' ' C r a n k Angle ( d e g ) '';(>')
426 7-.7 FORMAT* I X , ' X T I T ' ' V o l u m e ( c u . i n . )'';&')
427 "0= FORMAT*ix,'YTIT ' ' P r e s s u r e p s i '';&')
4 28 709 FORMAT*IX,'YG;XG')
429 END
 157

APPENDIX D - COMPUTER PROGRAM FOR DATA PROCESSING

 158

Listing of APP.PROG 1 a t 22:58:59 on APR 6, 1984 f o r C C i d = AFPH

1 C
2 C
3 C T h i s program p r o c e s s e s data from d i e s e l engine
4 C
5 C
6 EXTERNAL WTQIO
7 EXTERNAL GETADR
8 EXTERNAL ASNLUN
9 C
10 INTEGER LISTC200), IDATC200), IPARM(6), I STAT(2)
1 1 INTEGER YES, NO, ANS, IACTIV(2)
12 REAL LOAD
13 c
14 YES = 1
15 NO = 0
16 SCALE = 32768.0
1 7 c
18 c calibration constants f o r gas flow
19 c
20 C1GAS = 2 . 2 2
21 C2GAS = -0.0194
22 c
23 CALL PERFRM(VOLDSL,VOLGAS,VOLAIR,SPEED,LOAD,QINPPW,BM;
24 1 POWER,THRMEF,VOLEFF,PERDSL,RAF,RAD,RAG,0)
25 c
26 WRITE(5,248)
27 READ(5,150) VLOAD
28 LOAD = 5.0 * VLOAD
29 WRITE(5,249)
30 READ(5,150) SPEED
31 WRITE(5,247)
32 READ(5,150) DPQAIR
33 VOLAIR = 2.173 + 0.221 * DPQAIR
34 WRITE(5,245)
35 READ(5,150) DPQGAS
36 c c o n v e r t p a s c a l t o i n water
37 VOLGAS = DPQGAS / 248.8
38 VOLGAS = CI GAS * VOLGAS + C2GAS * VOLGAS * VOLGAS
39 WRITE(5,246)
40 READ(5,150) VOLDSL
41 VOLDSL = VOLDSL * 60.0
42 CALL PERFRM(VOLDSL,VOLGAS,VOLAIR,SPEED,LOAD,QINPPW,BMEP
43 1 POWER,THRMEF,VOLEFF,PERDSL,RAF,RAD,RAG,1)
44 WRITE(5,250)
45 WRITE(5,251) SPEED, LOAD, VOLAIR, VOLDSL
46 WRITE(5,254)
47 WRITE(5,255) QINPPW, POWER, BMEP, THRMEF, VOLEFF
48 WRITE(5,256)
49 WRITE(5,257) VOLGAS, PERDSL
50 WRITE(5,258)
51 WRITE(5,259) RAF, RAD, RAG
52 c save data i n a f i l e ?
53 WRITE(5,224)
54 READ(5,120) ANS
55 IF (ANS .EQ. NO) GO TO 70
56 IF (L .EQ. 1) CALL ASSIGN(2, 'OUT.DAT')
57 WRITE(2,250)
58 WRITE(2,251) SPEED, LOAD, VOLAIR, VOLDSL
 159

Listing of APP.PROG 1 a t 22:58: 59 on APR 6, 1984 f o r C C i d = AFPH

59 WRITE(2,252)
60 WRITE(2,253) DPTURB, DPCOMP, T1TURB, T2TURB, T1COMP, T2COMP
61 WRITE(2,254)
62 WRITE(2,255) QINPPW, POWER, BMEP, THRMEF, VOLEFF
63 WRITE(2,256)
64 WRITE(2,257) VOLGAS, PERDSL
65 WRITE(2,258)
66 WRITE(2,259) RAF, RAD, RAG
67 C
68 70 CONTINUE
69 WRITE(5,226)
70 READ(5,120) ANS
71 IF (ANS .EQ. NO) GO TO 999
72 500 CONTINUE
73 C
74 C
75 999 CONTINUE
76 STOP
77 C
78 100 FORMAT(1X,F16.5)
79 101 FORMAT(1X, I 2)
80 1 02 FORMAT(IX, 12)
81 103 FORMAT(05)
82 110 FORMAT(A5)
83 120 FORMAT(II)
84 150 FORMAT(F12.4)
85 C
86 C
87 224 FORMAT(IX,' ???? E n t e r 1 or 0 t o save d a t a ! ' , $ )
88 225 FORMAT(1X,I 5,2X,E14.6)
89 226 FORMAT(//,'???? To r e r u n e n t e r 1-or 0',$)
90 245 FORMAT(IX,' >> > E n t e r Gas Flow i n P a s c a l :',$)
91 246 FORMAT(IX,' >> > E n t e r D i e s e l Flow i n l i t r e / m i n : ' , $ )
92 247 FORMAT(IX,' >> > E n t e r A i r Flow i n P a s c a l : ' , $ )
93 248 FORMAT(1X,' >> > E n t e r Load i n V o l t a g e : ' , $ )
94 249 FORMAT(IX,' >> > E n t e r E n g i n e Speed i n rpm:',?)
95 250 FORMAT(1X,' Speed (rpm) Load ( l b ) A i r Flow (ft3/min) ',
96 1 ' D i e s e l Flow ( l t r / h r ) ' )
97 251 FORMAT(3X,F7.2,7X,F7.3,7X,F10.3,5X,F10.4)
98 253 FORMAT(1X,6(F8.2,2X))
99 254 FORMAT(5X,'Heat cons Power out BMEP Therm e f f V o l e f f ',
100 1 5X,'(BTU/hr-hp) (hp) (psi) (%) (%)
101 255 FORMAT(1X,F15.2,2X,F8.3,2X,F8.3,2X,F8.4,2X,F8.4)
102 256 FORMAT(10X,'Gas Flow d i e s e l input p r o p o r t i o n (heat)',/,
103 1 10X,'(ft3/min) ( p e r c e n t t o t a l h e a t ) ')
104 257 FORMAT(1 OX,F12.3,5X,F8.2,' %')
105 258 FORMAT(10X,' LAMDA ( t o t ) LAMDA ( d s l ) LAMDA ( g a s ) ' )
106 259 FORMAT(1 OX, 3F16.2)
107 END
108 C
109 C
110 C
111 SUBROUTINE PERFRM(VOLDSL,VOLGAS,VOLAIR,SPEED,LOAD,QINPPW,
112 1 BMEP,POWER,THRMEF,VOLEFF,PERDSL,RAF,RAD,RAG,INDEX)
113 C
114 C computes p e r f o r m a n c e characteristics
115 C
116 REAL LOAD,LHVDSL,LHVGAS
 160

;ting of A P P . PROG 1 a t 22:58:59 o n APR 6, 1984 f o r C C i d = A F P H

1 17 C
1 IB IF (INDEX .NE. 0) GO TO 10
1 19 C
1 20 C initialize const values
121 C
1 22 V1SCAG = 1.669
123 DENDSL = 0.8697 * 6 2 . 2 7
124 DENGAS = 0.044386
125 DENAIR = 0.07541
126 STCDSL = 15.0
127 HHVDSL = 1058288.4
128 STCGAS •= 16.7
129 LHVDSL = 1002560.0
1 30 HHVGAS = 1024.7
131 LHVGAS = 926.0
132 DSPLMT = 425.04
133 ARMLEN = 17.5 / 12.0
1 34 C
135 C conv. factors
136 C
137 F3PLTR = 0.03531
138 HPPFPM = 1 . 0 / 33000.
1 39 BTUPHP = 2 5 4 4 . 4 3 3 / 1.01387 / 0.986315
140 PI = 3.1415
141 C
142 RETURN
143 C
144 10 CONTINUE
145 c
146 c process data
147 c
148 TQDYNO • LOAD * ARMLEN
149 POWER = TQDYNO * S P E E D * 2.0 * PI * HPPFPM
1 50 VSWEPT = DSPLMT * S P E E D / 2.0
151 BMEP = (POWER / HPPFPM * 12.0) / VSWEPT
152 DSLFLW = VOLDSL * F3PLTR
153 GASFLW = VOLGAS * V 1 S C A G
154 FMAIR = DENAIR * V O L A I R
155 FMDSL = DENDSL * DSLFLW / 6 0 . 0
156 FMGAS = DENGAS * GASFLW
157 RAD = FMAIR / ( S T C D S L * F M D S L )
158 RAG = 0.0
159 IF (FMGAS .GT. 0.0) RAG = FMAIR / ( S T C G A S * FMGAS)
160 RAF = FMAIR / ( S T C D S L * FMDSL + STCGAS * FMGAS)
161 VOLEFF = V O L A I R / (VSWEPT / 12.0 / 12.0 / 12.0) * 100. 0
162 PERDSL = LHVDSL*DSLFLW / (LHVDSL*DSLFLW+LHVGAS*GASFLW* 60
163 PERDSL •= P E R D S L * 1 0 0 . 0
164 G
165 C b r a n c hi o u t n o - l o a d ie. idling
166 C
167 IF (ABS(LOAD) .LT. 0.00001) G O T O 11
168 QINPPW = ( L H V D S L * D S L F L W + LHVGAS*GASFLW* 6 0 . 0 ) / POWER
169 THRMEF = BTUPHP / QINPPW * 1 0 0 . 0
170 RETURN
171 C
172 1 1 CONTINUE
173 c
174 QINPPW = 0.0
 161

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175 THEMEF = 0.0

176 C
177 RETURN
178 END
 162

APPENDIX E - COMPUTER PROGRAM FOR APPARENT ENERGY RELEASE

 163

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1 C
2 C
3 C
4 C . Rate of Heat R e l e a s e A n a l y s i s Program
5 c
6 c w r i t t e n by : Seaho Song
7 c Dec / 1983
8 c
9 c
10 c
1 1 c
12 c T h i s program r e a d s i n t h e c y l i n d e r p r e s s u r e data
13 c t o compute t h e r a t e o f heat r e l e a s e f o r e v e r y C A .
14 c
15 c
16 c
17 c
18 c
19 c
20 Cmmmnmmmmmmmmmmmnmmmnunmmmmmm
21 Cm m
22 Cm Main Routine m
23 Cm m
24 CiTunmmmmmmmrnmmmnimnunmmmmmmmmmmmmmnuninmrnrnmmmmmmmmrnmmnirnmmminrnmrn
25 C
26 C
27 IMPLICIT REAL*8(A-H,0-Z)
28 REAL*8 CYLVOL(180)
29 REAL* 8 GAS ( 1 0 ) , GASNEWOO), P(180)
30 REAL*8 HEATRT(180),ANGLE(180), GASNE0(10),DH01(10),
31 REAL*8 F ( 2 ) , D F D X ( 2 , 2 ) , X ( 2 ) , D X ( 2 ) , F E P ( 2 ) , X E P ( 2 ) ,
32 1 WORKAR(2,2)
33 REAL*8 NTOT, SAVGAS(180,10)
34 INTEGER I PERM(4)
35 COMMON / GEOM / ARM, ROD, BORE, STROKE, VCLEAR
36 COMMON / EXPMT/ SPEED, BMEP
37 COMMON / PROP 1/ DENAIR, DENDSL, DENNG, WTDSL, WTNG,
38 1 WTAIR
39 COMMON /THDYPR/ H0F(10), R0, WT(10), NGAS
40 c
41 c
42 c specify cylinder geometry i n inches, then convert
43 c into metric.
44 c
45 CONVF1 = 0.0254
46 ARM =3.0 * CONVF1
47 ROD = 9.595 * CONVF1
48 BORE =4.75 * CONVF1
49 STROKE =6.0 * CONVF1
50 VCLEAR = 6.444 * CONVF1**3
51 c
52 c To smoothen p r e s s u r e d a t a SET I SOOT = 1
53 c To use heat t r a n s f e r model SET IHTRSF = 1
54 c To consider d i s s o c i a t i o n SET IDSSOC = 1
55 c To o b t a i n output a p p r o p r i a t e
56 c f o r p l o t t i n g SET IPLOT = 1
57 c
58 c Setting ISMOOT = 0; IHTRSF = 0; IDSSOC'= 0 w i l l
 164

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59 C assume unsmoothed, a d i a b a t i c w i t h no d i s s o c i a t i o n .
60 C
61 ISMOOT = 1
62 IHTRSF = 1
63 IDSSOC = 1
64 I PLOT = 1
65 c
66 EPSIL = 0.1E-1
67 c
68 c compute c y l i n d e r volume a t e v e r y CA.
69 c
70 CALL GEOMTR(CYLVOL, ANGLE)
71 c
72 c a s s i g n c o e f f i c i e n t s f o r thermodynamic
73 c p r o p e r t i e s of v a r i o u s g a s e s .
74 c
75 CALL READPR
76 c
77 c read i n c y l i n d e r p r e s s u r e d a t a , i n j e c t e d d i e s e l
78 c amount, CH4 amount, and i n j e c t i o n & i g n i t i o n
79 c characteristics.
80 c
81 CALL DATAIN(GAS,P,DSLAMT,INJBEG,INJEND,IGNBEG)
82 c
83 c w r i t e out t h e mode of o p e r a t i o n and bmep i n kPa
84 c
85 BMEPKP = BMEP * 6.8 95D0
B6 IF (GAS(2) .LT. 0.1D-12)
87 1 WRITE(6,210) SPEED, BMEPKP
88 IF (GAS(2) .GE. 0.1D-12)
89 2 WRITE(6,211) SPEED, BMEPKP
90 c
91 c smooth t h e P d a t a
92 c
93 IF (ISMOOT .NE. 1) GO TO 10
94 c
95 CALL SMOOTP(P, ANGLE, IGNBEG, IPOK)
96 IF (IPOK .EQ. 0) WRITE(6,914)
97 IF (IPOK .EQ. 0) STOP
98 c
99 10 CONTINUE
100 c
101 c set up f o r i n i t i a l stage
102 c
103 c FRREM keeps t r a c k of f r a c t i o n of f u e l remaining.
1 04 c FRBURN .. .. burnt.
1 05 c
1 06 c The subscript 1 r e f e r s t o the p r e v i o u s step
107 c 2 .. present
108 c
109 FRREM = 1.0
110 FRBURN = 0.0
1 1 1 QCACCM = 0.0
112 PI . = P(1)
1 13 V1 = CYLVOL(1)
1 14 Tl = P1 * VI / R0
115 c
116 c .TOTMAS - t o t a l mass o f gases p r e s e n t
 165

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1 17 C NTOT - total number of Kmoles of gases p r e s e n t

118 c
119 c GASM-10) - # of Kmoles of each gas p r e s e n t
120 c GASNEW(I-IO) - used t o update GAS(I-IO)
121 c
122 TOTMAS = 0.D0
123 NTOT = 0.D0
124 DO 29 J=1,NGAS
125 GASNEW(J) = GAS(J)
126 TOTMAS = TOTMAS + GAS(J) * WT(J)
127 NTOT = NTOT + GAS(J)
128 29 CONTINUE
129 C
130 C DH0TOT - E n t h a l p y a t T1 minus the E n t h a l p y a t 25
131 C f o r t h e t o t a l gas
132 C
133 CALL DH0FN(T1,DH01)
1 34 DH0TOT = 0.0
135 DO 31 1=1,NGAS
136 DH0TOT = DH0TOT + GASNEW(I) * DH01(I)
137 31 CONTINUE
138 U2 = DH0TOT - NTOT * R0 * T1
139 C
140 C w r i t e headings f o r the output
141 c
1 42 WRITE(6,200)
143 c
144 c c a l c u l a t i o n of r a t e of heat r e l e a s e i s carried
145 c out f o r each C A . d e g r e e .
146 c
147 DO 50 ITH=2,180
148 P2 = P(ITH)
1 49 V2 = CYLVOL(ITH)
150 c
151 c update t h e number of Kmoles of d i e s e l
152 c injected.
153 c
1 54 IF (ITH .EQ. INJBEG)
1 55 1 GAS(1) = GAS(1) + DSLAMT
156 IF (ITH .NE. INJBEG) GOTO 750
157 c
158 TOTMAS = 0.D0
159 NTOT = 0.D0
160 DO 732 J=1,NGAS
161 TOTMAS = TOTMAS + GAS(J)*WT(J)
162 NTOT = NTOT + GAS(J)
163 732 CONTINUE
164 750 CONTINUE
165 IF (ITH .GE. INJBEG) GO TO 55
166 C
167 c no c o m b u s t i o n .
168 c. processes compression stroke.
169 c
170 25 CONTINUE
171 FRAC = 0.0
172 RN = NTOT * R0
173 T2 = P2 * V2 / RN
174 C
 166

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175 C compute r a t e of heat transfer, internal energy.

176 C
177 CALL UPROD(P1,P2,T1,T2,V1,V2,GAS,GASNEW,DHO1,FRAC,
178 1 DHO,NTOT,TOTMAS,U2RES,QC,QHT,0,1HTRSF)
179 GO TO 58
180 C
181 C combustion i s t a k i n g p l a c e ,
182 C p r o c e s s e s combustion and e x p a n s i o n stroke.
183 C
184 c55 CONTINUE
185 C
186 c a s s i g n i n i t i a l guess v a l u e s f o r f r a c t i o n of
187 c f u e l b u r n t and t h e gas m i x t u r e t e m p e r a t u r e .
188 c
189 IF (FRAC .LT.0.1D-20) FRAC 0.1D-6
190 X(1) = FRAC
191 X(2) = P2 * V2 / R0 / NTOT
192 c
1 93 c u s e i n g m o d i f i e d Newton's method, t h e f o l l o w i n g
1 94 c two a r e c a l c u l a t e d i t e r a t i v e l y :
195 c FRAC - f r a c t i o n of f u e l b u r n t
196 c T2 - gas m i x t u r e t e m p e r a t u r e
197 c
198 c
199 c the following system of two n o m l i n e a r e q a t i o n s
200 c are solved
201 c
202 c P - (nRT/V)
203 c 2 2
204 c U - U + work - Qhtr
205 c 2 11 2 1 2
206 c
207 c for
208 c
209 c X FRAC
210 c T2
21 1 c
212 DO 650 L650=1,50
213 c
214 c g i v e n X, compute F
215 c
216 CALL GETF(X;F,GAS,T1,P1,P2,V1,V2,NTOT,
217 GASNEW,QHT,QC,TOTMAS,DH 01,DH 0,1HTRSF,IDS SOC)
218 c
219 c if solution i s found, terminate the i t e r a t i o n .
220 c
221 IF ((DABS(F(1)).LT.1.0).AND.(DABS(F(2)).LT.0.1E-4))
222 GO TO 59
223 c
224 c formulate the J a c o b i a n matrix of F a s :
225 c
226 c
227 c dF / dX dF / dX
228 c dF/dX 1 1 1 2
229 c
230 c dF / dX dF / dX
231 c 2 1 2 2
232 c
 167

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233 C
234 DO 651 LJ=1,2
235 DO 652 LI=1,2
236 XEP(LI) = X ( L I )
237 652 CONTINUE
238 X E P ( L J ) = X E P ( L J ) + EPSIL
239 CALL GETF(XEP,FEP,GAS,T1,P1,P2,V1,V2,NTOT,
240 1 GASNEW,QHT,QC,TOTMAS,DH01,DH0,IHTRSF,IDSSOC)
241 DO 653 LI=1,2
242 DFDX(LI,LJ) = ( F E P ( L I ) - F ( L I ) ) / E P S I L
243 653 CONTINUE
244 651 CONTINUE
245 C
246 C the iteration scheme i s as f o l l o w s :
247 C
248 C X = X + DX
249 C i+1 i
250, c
251 c DX i s o b t a i n e d by s o l v i n g
252 c
253 c (dFdX) DX = -F
254 c i i
255 c
256 c
257 F(1) = -F( 1 )
258 F(2) = -F(2)
259 c
260 c the r o u t i n e SLE i a a UBC L i b r a r y s u b r o u t i n e
261 c which s o l v e s a system of l i n e a r e q u a t i o n s .
262 c
263 CALL SLE(2,2,DFDX,1,2,F,DX,I PERM,2,WORKAR,
264 1 DET,JEXP)
265 c
266 DO 654 LI=1,2
267 X(LI) = X ( L I ) + DX(LI)
268 654 CONTINUE
269 650 CONTINUE
270 66 CONTINUE
271 C
272 C iteration failed, terminate the execution.
273 C
274 WRITE(6,913) L650, X ( 1 ) , X ( 2 )
275 GO TO 850
276 c
277 c solution found
278 c
279 59 CONTINUE
280 FRAC = X ( 1 )
281 T2 = X(2)
282 58 CONTINUE
263 C
284 C compute t h e a c c u m u l a t e d h e a t r e l e a s e ,
285 C a c c u m u l a t e d f r a c t i o n of f u e l b u r n t .
286 C
287 QCACCM = QCACCM + QC
288 IANGLE = ITH - 90
289 FRBURN = FRBURN + FRAC * FRREM
290 FRREM = 1 . 0 - FRBURN
 168

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291 C
292 C w r i t e out the s o l u t i o n f o r the c u r r e n t CA.
293 C
294 WRITE(6,201) IANGLE, P2, T2, QHT, QC, QCACCM,
295 1 FRAC, FRBURN, V2
296 C
297 C shift index f o r next C A . computation.
298 C
299 T1 = T2
300 P1 = P2
301 V1 = V2
302 DO 71 1=1,NGAS
303 DH01(I) = DHO(I)
304 71 CONTINUE
305 IF (ITH .LT. 80) GO TO 851
306 IF (ITH .GT. 170) GO TO 851
307 C
308 C w r i t e out the s o l u t i o n for plotting purpose.
309 C
310 IF (IPLOT .NE. 1) GO TO 851
311 WRITE(1,205) ANGLE(ITH), P2
312 WRITE(2,205) ANGLE(ITH), T2
313 WR1TE(3,205) ANGLE(ITH), QCACCM
314 WRITE(4,205) ANGLE(ITH), QC
315 851 CONTINUE
316 C
317 C save the i n s t a n t a n e o u s gas m i x t u r e c o m p o s i t i o n
318 C t o w r i t e out a t the end of the r a t e of heat r e l e a s e
319 C output.
320 C
321 IF (ITH .LT. ( I N J B E G - 1 ) ) GO TO 50
322 DO 61 J=1,NGAS
323 GAS(J) = GASNEW(J)
324 SAVGAS(ITH,J) = GAS(J)
325 61 CONTINUE
326 C
327 C t h i s i s t h e end of t h e p r o c e s s f o r one C.A..
328 C C A . i s i n c r e m e n t e d end the p r o c e s s proceeds
329 C t o the next C A . .
330 C
331 50 CONTINUE
332 850 CONTINUE
333 C
334 C end of a l l the p r o c e s s e s e .
335 C w r i t e out the gas m i x t u r e c o m p o s i t i o n f o r each C A .
336 C from the C A . j u s t p r i o r t o the d i e s e l i n j e c t i o n .
337 C
338 WRITE(6,202)
339 INJBM1 = INJBEG - 1
340 DO 72 ITH=INJBM1,180
341 IANGLE = ITH - 90
342 WRITE(6,203) IANGLE, (SAVGAS(ITH,J), J=1,NGAS)
343 72 CONTINUE
344 STOP
345 200 FORMAT('- CA P kPa T deg K, Q htr.(kJ), Q r'
346 1 'elease , Q accum Frac F r a c cum Vol')
347 201 FORMAT ( I X , 1 5 , 1 O E M . 5 )
348 202 FORMAT( IGas Comp(Kmol) D s l
1
CH4 N2 02
 169

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349 1 ' C02 H20 H2 OH CO

350 2 NO ' )
351 203 FORMAT(1X,I 3,7X,10E10.4)
352 205 FORMAT(1X,2E14.6)
353 210 FORMATCl S-D O p e r a t i o n Speed = ',F6.1,' rpm, ',
354 1 ' bmep = ',F5.1,' kPa ')
355 211 FORMATCl D-F O p e r a t i o n Speed = ',F6.1,' rpm, ',
356 1 ' bmep = *,F5.1,' kPa ' )
357 913 FORMATC- NO Conv i n fr&T L, f r , T = ',I5,2E14.5)
358 914 FORMATC- F a i l e d t o Smooth P data ')
359 END
360 C
361 Cssssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
362 c s
363 SUBROUTINE READPR
364 c s
365 Cssssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
366 c s
367 c T h i s r o u t i n e a s s i g n s v a l u s e , f o r v a r i o u s gases, s
368 c d e n s i t y , m o l e c u l a r w e i g h t , e n t h a l p y of f o r m a t i o n , s
369 c number of d i f f e r e n t k i n d of gases c o n s i d e r e d , s
370 c i d e a l gas c o n s t a n t . s
371 c s
372 Cssssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
373 c
374 c
375 IMPLICIT REAL*8(A-H,0-2)
376 COMMON /PROP 1/ DENAIR,DENDSL,DENNG,WTDSL,WTNG,WTAIR
377 COMMON /THDYPR/ H0F(10),RO,WT(10),NGAS
378 c
379 c density
380 c
381 DENAIR = 0.337600E-01
382 DENDSL = 0.848900E+00
383 DENNG = 0.1B5760E-01
384 c
385 c molecular weight o f i n t a k e gases
386 c
387 WTDSL = 0.170000E+03
388 WTNG = 0.160000E+02
389 WTAIR = 0.137280E+03
390 c
391 c number o f d i f f e r e n t k i n d o f gases c o n s i d e r e d , and
392 c the i d e a l gas c o n s t a n t .
393 c
394 NGAS = 10
395 R0 = 0.831425E+01
396 c
397 c enthalpy of formation a t standard c o n d i t i o n .
398 c
399 H0F(1) = -.290871E+06
400 H0F(2) = -.748730E+05
401 H0F(3) = 0.000000E+00
402 H0F(4) = 0.000000E+00
403 H0F(5) = -.393522E+06
404 H0F(6) = -.241827E+06
405 H0F(7) = 0.000000E+00
406 H0F(8) = 0.394630E+05
 170

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407 H0FC9) = -.110529E+06

408 H0F(10)= 0.905920E+05
409 C
410 C molecular weight
41 1 C
412 WT( 1) 0. 170000E+03
413 WT(2) 0.160400E+02
414 WT(3) 0.280130E+02
41 5 WT(4) 0.319990E+02
416 WT(5) 0.440100E+02
417 WT(6) 0.180150E+02
418 WT(7) 0.201600E+01
419 WT(8) 0.170070E+02
420 WT(9) 0.280100E+02
421 WT(10) 0.460000E+02
422 C
423 C
424 RETURN
425 END
426
427 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
428 C s
429 SUBROUTINE GEOMTR(CYLVOL, ANGLE)
430 C s
431 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
432 C s
433 C T h i s r o u t i n e a s s i g n s v a l u s e f o r the engine geometry s
434 C and computes c y l i n d e r volume a t each C A . s
435 C s
436 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
437 c
438 IMPLICIT REAL*8(A-H,0-Z)
439
440 COMMON / GEOM / ARM, ROD, BORE, STROKE, VCLEAR
441 REAL*B C Y L V O L O 8 0 ) , ANGLE(180)
442 c
443 c Computes V, ANGLE f o r Theta=-90,89 (Crank Angle)
444 c
445 c in c u . meter
446 c
447 XAREA 3.14* BORE * BORE / 4.0
448 ARMSQ ARM * ARM
449 RODSQ ROD * ROD
450 RADPDG 3.14 / 180.0
451
452 compute cylinder volume
453
454 ITH1 = -89
455 ITH2 = 90
456 DO 100 ITH=ITH1,ITH2
457 TH = DFLOAT(ITH)
458 RD = TH * RADPDG
459 X = ROD+ARM*(1.O-DCOS(RD))-DSQRT(RODSQ-ARMSQ*DSIN(RD)**2)
460 CYLVOL(ITH+90) = XAREA * X + VCLEAR
461 ANGLE(ITH+90) = TH
462 100 CONTINUE
463 C
464 C
 171

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465 RETURN
466 END
467 C
468 Cssssssssssssssssssssssssssssssssssssssss5sssssssssssssssssss
469 c S
470 SUBROUTINE DATAIN(GAS, P, DSLAMT,INJBEG,INJEND,IGNBEG)
471 c s
472 CSSSSSSSSS5SSSSSSSSSSSSSSSSSSSSSSSSSSSSS5SSSSSSSSSSSSSSSSS5SS
473 c S
474 c T h i s r o u t i n e r e a d s i n t h e d a t a f o r i n j e c t i o n and s
475 c i g n i t i o n c h a r a c t e r i s t i c s , c y l i n d e r p r e s s u r e , engine s
476 c speed, BMEP, flow r a t e s of a i r , g a s , d i e s e l . s
477 c The flow r a t e s a r e c o n v e r t e d t o number of Kmoles s
478 c per c y c l e r . s
479 c s
480 CSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSS5SS5
481 c
482 IMPLICIT REAL*8(A-H,0-Z)
483 c
484 COMMON / EXPMT / SPEED, BMEP
485 COMMON / PROP 1 / DENAIR, DENDSL, DENNG, WTDSL,
486 1 VJTNG, WTAIR
487 REAL*8 GAS(10), P(180)
488 c
489 c Reads i n from Log U n i t 10
490 c
491 c SPEED - engine speed i n rpm
492 c BMEP - load i n p s i
493 c
494 c QAIR - a i r flow i n f t 3 / m i n
495 c QDSL - d i e s e l flow i n l t r / h r
496 c QNG - n a t gas flow i n f t 3 / m i n
497 c
498 c INJBEG - b e g i n n i n g of d i e s l i n j i n deg C A .
499 c INJEND - end of d i e s e l i n j i n deg C A .
500 c IGNBEG - b e g i n n i n g of i g n i t i o n i n deg C A .
501 c"
502 c P(1-180) - cylinder pressure in psi
503 c
504 c
505 READ(10,100) SPEED, BMEP
506 READ(10,101) QAIR, QDSL, QNG
507 READ(10,102) INJBEG,INJEND,IGNBEG
508 c
509 JA = 1
510 DO 10 L=1,36
51 1 JB = JA + 4
512 READ(10,103) ( P ( J ) , J=JA,JB)
513 JA = L * 5 + 1
514 10 CONTINUE
515 c
516 c c o n v e r t p r e s s u r e data from p s i t o kPa
517 c
518 DO 20 J=1,180
519 P ( J ) = P ( J ) * 6.895
520 20 CONTINUE
521 c
522 c compute # o f moles p e r c y c l e :
 172

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523 C
524 C GAS(2) - n a t gas
525 c (3) - N2
526 c (4) - 02
527 c
528 DO 30 L=1,20
529 GAS(L) = 0.0
530 30 CONTINUE
531 c
532 c making sure of # o f c y l i n d e r = 4
533 c
534 GAS(2) = QNG * DENNG / SPEED * 2.0 / WTNG / 4.0
535 AIRMOL = QAIR * DENAIR / SPEED * 2.0 / WTAIR
536 GAS(3) = 3.76 * AIRMOL / 4.0
537 GAS(4) = AIRMOL /4.0
538 c
539 c compute amount of d i e s e l injected i n kmoles
540 c
541 DSLAMT = QDSL / 60.0 * DENDSL / SPEED * 2.0 / WTDSL / 4.1
542 c
543 c
544 RETURN
545 100 FORMAT(1X, F6.1, 1X, F5.1)
546 101 FORMAT(IX, F5.1, 1X, F5.2, 1X, F5.2)
547 102 FORMAT(IX,3(13,IX))
548 103 FORMAT(IX, 5(F6.1,1X))
549 END
550 c
551 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
552 c s
553 SUBROUTINE UPROD(P1,P2,T1,T2,V1,V2,GAS,GASNEW,DHO1,FRAC,
554 1 DHO,NTOT,TOTMAS,U2RES,QC,QHT,ICOMB,IHTRSF)
555 c s
556 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
557 c s
558 c T h i s r o u t i n e c h e c k s out whether t h e f i r s t law i s met. s
559 c The d e v i a n c e from t h e f i r s t law i s d e s i g n a t e d by U2RES.S
560 c s
561 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
562 c
563 IMPLICIT REAL*8(A-H,0-Z)
564 c
565 REAL*8 GAS(10),GASNEW(10),R1(10),DH01(10),DH0(10),NTOT
566 COMMON /THDYPR/ H 0 F ( 1 0 ) , R0, WT(10), NGAS
567 c
568 WORK = 0.5D0 * (PI + P2) * <V2 - V1)
569 c
570 CALL DH0FN(T2,DH0)
571 c
572 IF (ICOMB .EQ. 0) GOTO 5
573 c
574 c compute t h e heat r e l e a s e due t o combustion during
575 c the c u r r e n t C A . i n t e r v a l .
576 c
577 QC = O.ODO
578 DO 20 1=1,NGAS
579 QC = QC + (GASNEW(I) - G A S ( I ) )
580 1 * (H0F(I) + DH0(I) - R0 * T2)
 173

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581 20 CONTINUE
582 C
583 QC = -QC
584 C
585 5 CONTINUE
586 IF (I COMB .EQ. 0) QC = 0.0
587 C
588 C
589 DU = 0.D0
590 R0DT = R0 * (T2 - T l )
591 DO 30 1=1,NGAS
592 DU = DU + GAS(I) * (DHO(I) - DHO1(1) - R0DT)
593 30 CONTINUE
594 C
595 C comput heat t r a s f e r .
596 C i f IHTRSF i s s e t t o 0, a d i a b a t i c processe i s
597 C assumed.
59B C
599 QHT = 0.D0
600 IF (IHTRSF .NE. 1) GO TO 10
601 QHT = QHTRSF(T2,V2,GASNEW,TOTMAS)
602 C
603 10 CONTINUE
604 U2RES = DU + WORK - QC - QHT
605 c
606 RETURN
607 END
608 c
609 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
610 c s
61 1 SUBROUTINE STCHPD(GAS,FRAC,GASNEW)
612 c s
613 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
614 c s
615 c T h i s r o u t i n e computes the number of Kmoles of s
616 c s t o i c h i o m e t r i c combustion p r o d u c t , and y i e l d the s
617 c updated c o m p o s i t i o n of t h e gas m i x t u r e i n t h e s
618 c cylinder. s
619 c s
620 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
621 c
622 IMPLICIT REAL*8(A-H,0-2)
623 c
624 COMMON /THDYPR/ H O F ( l O ) , R0, WT(10), NGAS
625 REAL*8 GAS(10), GASNEW(lO), N, M
626 c
627 c M - number o f moles o f d i e s e l burnt at current C A .
628 c N - .. CH4
629 c
630 M = GAS(1) * FRAC
631 N = GAS(2) * FRAC
632 c
633 GASNEW(1) = GAS(1) - M
634 GASNEW(2) = GAS(2) - N
635 GASNEW(3) = GAS(3)
636 GASNEW(4) = GAS(4) - 18.5*M - 2.0*N
637 GASNEW(5) = GAS(5) + 12.0*M + N
638 GASNEW(6) = GAS(6) + 13.0*M + 2.0*N
 174

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639 DO 10 1=7,NGAS
640 GASNEW(I) = GAS(I)
641 10 CONTINUE
642 C
64 3 RETURN
644 END
645 C
646 Cssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
647 C s
648 SUBROUTINE SMOOTP(P, ANGLE, IGNBEG, IPOK)
649 C s
650 Cssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
651 C s
652 C T h i s r o u t i n e uses C u b i c - S p l i n e - L e a s t - S q u a r e s - F i t s
653 C t o smooth the c y l i n d e r p r e s s u r e d a t a . s
654 C The r a t e of p r e s s u r e r i s e i s n u m e r i c a l l y computed s
655 C from the UNSMOOTHED d a t a , and t h i s i s used i n s
656 C weight t o c o n t r o l the degree of l o c a l smoothness. s
657 C The w e i g h t i n g i s based on s c a t t e r n e s s t h e of s l o p o f s
658 C of the p r e s s u r e d a t a . s
659 C s
660 Cssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
661 C
662 IMPLICIT REAL*8(A-H,0-Z)
663 C
664 REAL* 8 P ( 1 8 0 ) , DPDTHO80), T O L ( l B O ) , ANGLE ( 1 80)
665 REAL*8 P D l ( l B O ) , PD2(180)
666 REAL*8 W(2000)
667 C
668 C compute dp/dTheta by f i t t i n g a q u a d r a t i c through
669 C 3 points
670 C
671 DO 5 J=2,179
672 DPDTH(J) = (P(J+1) - P ( J - l ) ) / 2.DO
67 3 5 CONTINUE
674 DPDTH(1) = (4.D0*P(2) -3.D0*P(1) - P ( 3 ) ) /2.D0
675 DPDTHC180) = (3.D0*P(180)-4.D0*P(179)+P(17B))/2.DO
676 C
677 C TOL0 c o n t r o l s the l o c a l smoothness.
678 C SVAL .. the g l o b a l
679 C
680 TOL0 = 1.0
681 SVAL = 1000.0
682 IPOK = 1
683 DO 10 1=4,177
684 JS = I - 3
685 JF = I + 2
686 AV = 0.0
687 DO 11 J=JS,JF
688 AV = AV + DPDTH(J)
689 11 CONTINUE
690 AV = AV / 10.0
691 C
692 C SD i s a measure of s c a t t e r n e s s i n t h e s l o p of
693 C of t h e p r e s s u r e d a t a . T h i s i s computed by
694 C c o n s i d e r i n g 4 adjacent p o i n t s .
695 C
696 SD = 0.0
 175

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697 DO 12 J=JS,JF
698 SD = SD + DABS(DPDTH(J) - AV)**2
699 12 CONTINUE
700 T O L ( I ) = TOL0 * DABS(SD / AV / AV)
701 C
702 C The degree of smoothness i s l e s s f o r c e d f o r
703 C the p o i n t s b e f o r the s t a r t of i g n i t i o n .
704 C
705 IF (DABS(DFLOATCI-IGNBEG)).LT.5.0) T O L ( I ) = T O L ( I ) / 1 0
706 10 CONTINUE
707 DO 20 1=1,3
708 TOL(I) = TOL(4)
709 20 CONTINUE
710 DO 30 1=178,180
71 1 TOL(I) = TOL(177)
712 30 CONTINUE
713 C
714 C The r o u t i n e s DSPLFT and DSPLN a r e UBC L i b r a r y
715 C s u b r o u t i n e s , which p e r f o r m s Least-Squares-Fit
716 C w i t h C u b i c - S p l i n e as b a s i s f u n c t i o n s .
717 C
718 CALL DSPLFT (ANGLE, P, TOL, SVAL, 180,W,5,613)
719 CALL DSPLN (ANGLE, P,PD1 ,PD2, 180,5.61 3)
720 RETURN
721 c
722 613 CONTINUE
723 c
724 c The f i t has failed
725 c
726 IPOK = 0
727 RETURN
728 END
729 c
730 CSSSSSSSS5SSSS5SS5SSSSSSSSSSSSSSSSSSSSSSS5SSSSSSSSSSSSSSS5
731 c s
732 SUBROUTINE VISCST(T,GAS,VISC)
733 c s
734 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssss
735 c s
736 c computes mean v i s c o u s i t y of gas m i x t u r e s . s
737 c s
738 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssss
739 c
740 IMPLICIT REAL*8(A-H,0-Z)
741 c
742 REAL*8 GAS(10)
743 COMMON /THDYPR/ H0F(10),R0,WT(10),NGAS
744 c
745 TM = T**0.645
746 VISC = GASO) * WT(1) * 1.33
747 VISC = VISC + GAS(2) * WT(2) * 3.35
748 VISC = VISC + GASO) * WT(3) * 4.57
749 VISC = VISC + GAS(4) * WT(4) * 5.09
750 VISC = VISC + GAS(5) * WT(5) * 3.71
751 VISC = VISC + GAS(6) * WT(6) * 3.26
752 c
753 TOTW = GAS(1)*WT(1)+GAS(2)*WT(2)+GAS(3)*WT(3)
754 1 + GAS(4)*WT(4)+GAS(5)*WT(5)+GAS(6)*WT(6)
 176

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755 C
756 VISC = VISC / TOTW * 10.**(-7) * TM
757 RETURN
758 END
759 C
760 C
761 C
762 Cffffffffffffffffffffffffffffffffffffffffffffffffffffffffffff
763 c f
764 DOUBLE PRECISION FUNCTION QHTRSF(T2,V2,GASNEW,TOTMAS)
765 c f
766 Cffffffffffffffffffffffffffffffffffffffffffffffffffffffffffff
767 c f
768 c T h i s r o u t i n e computes t h e heat t r a n s f e r a t c u r r e n t f
769 c Annand's model i s used. f
770 c f
771 Cffffffffffffffffffffffffffffffffffffffffffffffffffffffffffff
772 c
773 IMPLICIT REAL*8(A-H,0-Z)
774 c
775 COMMON / GEOM / ARM,ROD,BORE,STROKE,VCLEAR
776 COMMON /EXPMT / SPEED, BMEP
777 REAL*8 GASNEW(10)
778 c
779 c T h i s r o u t i n e employees Annand's model t o
780 c compute t h e r a t e of heat t r a n s f e r .
781 c
782 A = 0.47D0
783 C = 1.6D-12
784 CP = CP0VAL(T2,GASNEW) / TOTMAS
785 30 CONTINUE
786 PISVEL = SPEED * STROKE / 30.0
787 CALL VISCST(T2,GASNEW,VISC)
788 DENTOT = TOTMAS / V2
789 RENUM = DENTOT * PISVEL * BORE / VISC
790 REKD = CP * VISC / 0.7 / BORE * RENUM**(0.7)
791 c
792 c The w a l l t e m p e r a t u r e i s assumed t o be p r o p o r t i o n a l
793 c to the a p p l i e d l o a d .
794 c
795 TW = 0.484 * BMEP + 540.0
796 c
797 c The s u f a c e a r e o f t h e c y l i n d e r
798 c
799 SURFA = (V2 - VCLEAR) * 4.DO / BORE + 0.0304D0
800 QCONVC = A * SURFA * REKD* (T2 - TW)
801 QRAD = (1.6E-12)*(10.76)*C*SURFA*(T2**4-TW**4)
802 c
803 c
804 QHTRSF = -(QCONVC + QRAD) * (60./SPEED/360.)
805 c
806 c
807 c
808 RETURN
809 END
810 c
811 Cssssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
812 C S
 177

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813 SUBROUTINE DSSOCN(P,T,GAS1,GAS2,NTOT)

814 C S
B15 CS5SSSSSSSSSSSSSSSSSS5SSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSS5
816 c s
817 c T h i s r o u t i n e computes the e q u i l i b r i u m d i s s o c i a t i o n s
818 c p r o d u c t s . The t h e o r e t i c a l d e t a i l s i n c l u d i n g the s
819 c n u m e r i c a l methods a r e d e s c r i b e d i n the e x t e r n a l s
820 c documentation. s
821 c s
822 c The r e a c t i o n s c o n s i d e r e d are: s
823 c s
824 c 1. CO <==> CO + 1/2 0 s
825 c 2 2 s
826 c s
827 c 2. HO <==> 1/2 H + OH s
828 c 2 2 s
829 c 5
830 c 3. HO <==> H + 1/2 O s
831 c 2 2 2 s
832 c s
833 c 4. 1/2 N + 1 / 2 0 <==> NO s
834 c 2 2 s
835 c 5
836 CSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSSS5
837 c
838 IMPLICIT REAL*8(A~H,O-Z)
839 c
840 COMMON /THDYPR/ HOF(10),R0,WT(10),NGAS
841 REAL*8 GAS1(10), G A S 2 O 0 ) , NTOT
842 REAL*8 K ( 4 ) , F ( 5 ) , FEP(5),KPOP(4), DX(5), X ( 5 ) , XEP(5)
843 REAL*8 WORKDB(5,5), DXDBL(5), FDBL(5),DFDXDB(5,5),DETDBL
844 INTEGER I PERM(10)
845 c
846 SUMN = 0.0
847 DO 5 1=1,NGAS
848 GAS2(I) = GAS 1(I)
849 SUMN = SUMN + GAS 1(I)
850 5 CONTINUE
851 c
852 IF (GAS1(5) .GE. 0.1E-20) GO TO 500
853 NTOT = SUMN
854 RETURN
855 c
856 c
857 500 CONTINUE
858 c
859 c compute the e q u i l i b r i u m c o n s t a n t s f o r the g i v e n
860 c temperature.
861 c
862 K(1) = DEXP(DLOG(T)**(-7.4721)*(-0.65549E+8)+10.53)
863 K(2) = DEXP(DLOG(T)**(-7.0457)*(-0.30372E+8)+l0.159)
864 K(3) = DEXP(DLOG(T)**(-6.8674)*(-0.18879E+8)+8.7095)
865 K(4) = DEXP(DLOG(T)**(-7.3355)*(-0.16593E+8)+1.80127)
866 c
867 POP = 101.325DO/P
868 KP0P(1) = K(1) * K(1)*P0P
869 KP0P(2) = K(2) * K(2)*P0P
870 KP0P(3) = K(3) * K(3)*P0P
 178

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871 K P O P U ) = K(4) * K(4)

872 C
873 C if the r e a c t i o n s a r e i n s i g n i f i c a n t , then skip.
674 C
B75 RMK =0.0
876 DO 10 1=1,4
677 IF (DABS(KP0P(I ) ) .GT. RMK) RMK = DABSCKPOPdi
878 10 CONTINUE
879 IF (RMK .GT. 0.1E-5) GO TO 20
880 NTOT = SUMN
881 RETURN
882 c
883 c
884 20 CONTINUE
885 C
886 C initial guess
887 C .
888 X(1) = K(1)*DSQRT(P0P/SUMN/GAS1(4))*GAS1(5)
889 X(1) = X(1) - GAS 1(9) / SUMN
890 X(3) = K(3)*DSQRT(P0P/SUMN/GAS1(4))*GAS1(6)
891 X(3) = X(3) - GAS 1(7) / SUMN
892 X(2) = K(2)*DSQRT(P0P/SUMN)*GAS1(6)
893 X(2) = X(2)/DSQRT(DABS(GAS1(7)+X(3)*SUMN))
894 X(2) = X(2) - GAS1(8)/SUMN
895 X(4) = K(4)*DSQRT(GAS1(3)*GAS1(4))/SUMN
896 X(4) = X(4) - GAS 1 (10)/SUMN
897 X(5) = SUMN
898 40 CONTINUE
899 C
900 c S o l v e f o r X u s i n g m o d i f i e d Newton's method.
901 c The method i s p r e c i s e l y t h e same a s t h a t i n
902 c the main r o u t i n e , and t h e n o t a i o n s a r e a l s o
903 c n e a r l y t h e same.
904 c
905 TOL = DABS ( K P O P U ) ) * 0.1 D-4
906 DO 50 L=1,100 .
907 CALL EVALF(X,GAS2,KP0P,SUMN,F)
908 c
909 c I t i s s u f f i c i e n t t o check o n l y t h e F ( 4 ) ,
910 c s i n c e i t i s t h e most s l o w l y c o n v e r g i n g term.
91 1 c
912 IF (DABS(F(4)) .LT. TOL) GOTO 300
913 c
914 DO 51 LJ=1,5
915 EPSIL = DSQRT(0.1D-12+0.1D0*DABS(X(LJ)))
916 DO 52 LI=1,5
917 XEP(LI) = X(LI)
918 52 CONTINUE
919 X E P ( L J ) = X E P ( L J ) + EPSIL
920 CALL EVALF(XEP,GAS2,KP0P,SUMN,FEP)
921 DO 53 LI=1,5
922 DFDXDB(LI,LJ) = ( F E P ( L I ) - F ( L I ) ) / EPSIL
923 53 CONTINUE
924 51 CONTINUE
925 c
926 DO 55 LI=1,5
927 FDBL(LI) = -F(LI)
928 55 CONTINUE
 179

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929 C
930 C A g a i n , SLE i s a UBC L i b r a r y s u b r o u t i n e , which
931 C s o l v e s a system of l i n e a r e q u a t i o n s .
932 C In s o l v i n g t h e systems of e q u a t i o n s , t h e r o u t i n e
933 C r e t a i n s t h e decomposed m a t r i x .
934 C
935 CALL SLE(5,5,DFDXDB,1,5,FDBL,DXDBL,I PERM,5,WORKDB,
936 1 DETDBL,JEXB)
937 C
938 DO 54 LI=1,5
939 X ( L I ) = X ( L I ) + DXDBL(LI)
940 54 CONTINUE
94 1 NTOT = X ( 5 )
942 A = X(1) * NTOT
94 3 B = X(2) * NTOT
944 C = X(3) * NTOT
94 5 D = X ( 4 ) * NTOT
946 GAS2(3) = DABS(GAS1(3) - 0.5*D)
947 GAS2(4) = DABS(GAS1(4) + 0.5MA+C-D))
948 GAS2(5) = DABS(GAS1(5) - A)
949 GAS2(6) = DABS(GAS1(6) - B - C)
950 GAS2(7) = DABS(GAS1(7) + 0.5*B + C)
951 GAS2(8) = DABS(GAS1(8) + B)
952 GAS2(9) = DABS(GAS1(9) + A)
953 G A S 2 O 0 ) = DABS (GAS 1(10) + D)
954 C
955 C Here, once a J a c o b i a n m a t r i x i s formed f o r F,
956 C i t i s used f o r 3-4 i t e r a t i o n s , t h u s r e d u c i n g
957 C the c o s t .
958 C
959 DO 70 J70=1,3
960 CALL EVALF(X,GAS2,KP0P,SUMN,F)
961 DO 71 LI=1,5
962 FDBL(LI) = - F ( L I )
963 71 CONTINUE
964 C
965 C The r o u t i n e DBS i s a l s o a UBC L i b r a r y r o u t i n e .
966 C The r o u t i n e uses t h e decomposed m a t r i x by t h e
967 C r o u t i n e SLE t o v e r y e c o n o m i c a l l y compute new
968 C s o l u t i o n w i t h newly g i v e n FDBL
969 C
970 CALL DBS(5,1,5,FDBL,DXDBL,I PERM,5,WORKDB)
971 DO 73 LI=1,5
972 X ( L I ) = X ( L I ) + DXDBL(LI)
97 3 73 CONTINUE
974 C
975 C update t h e c o m p o s i t i o n of t h e gas m i x t u r e .
976 C
977 NTOT = X ( 5 )
978 A = X ( 1 ) * NTOT
979 B = X ( 2 ) * NTOT
980 C = X ( 3 ) * NTOT
981 D = X ( 4 ) * NTOT
982 GAS2(3) = DABS(GAS1(3) - 0.5*D)
983 GAS2(4) = DABS(GAS 1(4) + 0.5*(A+C-D))
984 GAS2(5) = DABS(GAS1(5) - A)
985 GAS2(6) = DABS(GAS1(6) - B - C)
986 GAS2(7) = DABS(GAS1(7) + 0.5*B + C)
 180

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987 GAS2(8) = DABS(GAS1(8) + B)

988 GAS2(9) = DABS(GAS1(9) + A)
989 G A S 2 O 0 ) = DABS (GAS 1(10) + D)
990 70 CONTINUE
991 50 CONTINUE
992 C
993 C I t e r a t i o n has f a i l e d . The e x e c u t i o n will terminate
994 c w i t h a p r o p e r e r r o r message.
995 c
996 WRITE(6,213)
997 STOP
998 c
999 c mission completed. Exit.
1 000 c
1001 300 CONTINUE
1002 RETURN
1003 213 FORMATC- xXxXxXxXxX F a i l t o Converge i n Dssocn xXxXxX
1 004 END
1005 c
1006 c
1007 Csssssssssssssssssssssssssssssssssssssssssssssssssssss
1008 c s
1009 SUBROUTINE EVALF(X,N,KP0P,SUMN,F)
1010 c S
1011 CSSSSSSSSSSSSSSSSSSSSSS5SSSSSSSSSSSSSSSSSSSSSSSS5SSSSS
1012 c s
1013 c T h i s r o u t i n e c h e c k s t h e a p p r o p r i a t e n e s s of s
1014 c the g i v e n p o s s i b l e s o l u t i o n f o r t h e s
1015 c equilibrium dissociation. The d e v i a t i o n i s s.
1016 c d e s i g n a t e d by t h e v e c t o r F. s
1017 c s
1018 CSSSSSSSSS5SSSSSSSSSSSSSSSSSSSSSSSSSSS5SSSSSSSSSSSSSSS
1019 c
1020 IMPLICIT REAL*8(A-H,0-Z)
1021 c
1022 REAL*8 X ( 5 ) , N ( 1 0 ) , KP0P(4), F ( 5 ) , NTOT
1 023 c
1024 NTOT = X(5)
1025 c
1026 TERM1 = N(4)/NTOT+0.5*(X(1)+X(3)-X(4))
1027 TERM2 = (N(7)/NTOT + 0.5*X(2) + X ( 3 ) )
1 028 TERM3 = (N(3)/NTOT-0.5*X(4))*(N(4)/NTOT
1029 1 +0.5*(X(1)+X(3)-X(4)))
1030 c
1031 F(1) = (N(9)/NTOT+X(1))**2*TERM1/(N(5)/NTOT-X(1))**2
1032 1 - KP0P(1)
1033 F(2) = TERM2*(N(B)/NTOT+X(2))**2/(N(6)/NTOT-X(2)-X(3))
1034 1 - KP0P(2)
1035 F(3) = TERM2**2 * TERM1
1036 1 / ( N ( 6 ) / N T O T - X ( 2 ) - X ( 3 ) ) * * 2 - KP0P(3)
1037 F(4) = (N(10)/NTOT+X(4))**2/TERM3
1038 1 -KP0P(4)
1039 F(5)- = (SUMN-NTOT)-0.5D0*(X(1)+X(2)+X(3))*NTOT
1 040 c
1041 c
1042 RETURN
1043 END
1044 c
 181

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1045 C
1046 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
1047 c s
1048 SUBROUTINE GETF(X,F,GAS,T1,P1,P2,V1,V2,NTOT,
1049 1 GASNEW,QHT,QC,TOTMAS,DHO1,DHO,IHTRSF,IDSSOC)
1050 c s
1051 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
1052 c s
1053 c T h i s r o u t i n e c h e c k s t h e a p p r o p r i a t e n e s s of t h e g i v e n s
1054 c p o s s i b l e s o l u t i o n s f o r t h e requrements f o r t h e f i r s t s
1055 c law and t h e i d e a l gas law. T h i s r o u t i n e i s used i n s
1056 c the main r o u t i n e f o r computing t h e f r a c t i o n of f u e l s
1057 c b u r n t and T2. s
1058 c s
1059 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
1060 c
1061 IMPLICIT REAL*8(A-H,0-Z)
1062 c
1063 REAL*8 X(2),F(2),GAS(10),GASNEW(10),DH0(10) ,DH01 (10)
1 064 REAL*8 NTOT,GASNE0(20)
1065 COMMON /THDYPR/ H0F(10),R0,WT(10),NGAS
1 066
1067
c
FRAC = X ( 1 )
1068 T2 = X(2)
1069 c
1 070 IF (IDSSOC .EQ. 0) GO TO 50
1071 CALL STCHPD(GAS,FRAC,GASNE0)
1072 CALL DSSOCN(P2,T2,GASNE0,GASNEW,NTOT)
1073 GO TO 51
1 074 50 CONTINUE
1075 CALL STCHPD(GAS,FRAC,GASNEW)
1076 51 CONTINUE
1 077 CALL UPROD(P1,P2,T1,T2,V1,V2,GAS.GASNEW,DHO1,FRAC,
1078 1 DHO,NTOT,TOTMAS,U2RES,QC,QHT,1,IHTRSF)
1 079 c
1 080 F(1) = P2 - NTOT * R0 * T2 / V2
1 081 F(2) = U2RES
1082 c
1083 c
1 084 RETURN
1085 END
1086 c
1087 Cffffffffffffffffffffffffffffffffffffffffffffffffffffffffff
1088 c f
1089 DOUBLE PRECISION FUNCTION CP0VAL(T,GAS)
1090 c f
1091 Cffffffffffffffffffffffffffffffffffffffffffffffffffffffffff
1092 c f
1093 c T h i s r o u t i n e computes t h e mean v a l u e of t h e s p e c i f i c f
1094 c heat Cp. f
1095 c f
1096 Cffffffffffffffffffffffffffffffffffffffffffffffffffffffffff
1097 c
1098 IMPLICIT REAL*8(A-H,0-Z)
1099 c
1 100 COMMON /THDYPR/ H 0 F ( 1 0 ) , R0, WT(10),NGAS
1101 REAL*8 GAS(10), C P 0 ( 1 0 ) , NTOT
1 102 c
 182

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103 C
104 TH = T / 100.0
105 TH2 = TH * TH
106 TQ = TH**(0.25)
107 TQ2 = TQ * TQ
108 TQ2 = TQ * TQ * TQ
109 TQ6 = TQ3*TQ3
110 C
1 1 1 CP0C1) = 104.18 + 465.5 * (T / 1000.0)
1 12 CP0(2) = -672.87 + 439.74*TQ - 24.875*TQ3 + 323.88/TQ2
1 13 CP0(3) = 39.06-512.79/TQ6+1072.7/TH2-820.4/(TH**3)
1 14 CP0(4) = 37.432 + 0.020102*TQ6-178.57/TQ6+236.88/TH2
115 CP0(5) = -3.7357+30.529*TQ2-4.1034*TH+0.024198*TH2
1 16 CP0(6) = 143.05-183.54*TQ+82.751*TQ2-3.6989*TH
1 17 CP0(7) = 56.505-702.74/TQ3+1165.O/TH-560.7/TQ6
118 CP0(8) = 81.546-59.35*TQ+17.329*TQ3-4.266*TH
1 19 CP0(9) = 69.145-0.70463*TQ3-200.77/TQ2+l76.76/TQ3
120 CP0(10)= 46.045+216.1/TQ2-363.66/TQ3+232.55/TH2
121 C
122 CPOVAL = 0.0
123 C DO 20 1=1,NGAS
124 DO 20 I=2,NGAS
125 CPOVAL = CPOVAL + C P 0 ( I ) * GAS(I)
126 20 CONTINUE
1 27 C
1 28 C
1 29 RETURN
1 30 END
131 C
132 C
133 C
134 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
135 c 5
136 SUBROUTINE DH0FN(T,DH0)
137 c s
138 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
1 39 c s
140 c T h i s r o u t i n e computes t h e change i n e n t h a l p y s
141 c between t h e g i v e n t e m p e r a t u r e and 298 deg K. s
142 c s
143 c The u n i t f o r DH0O-10) i s kJ/Kmol-K s
144 c s
145 Csssssssssssssssssssssssssssssssssssssssssssssssssssssssssss
146 c
147 IMPLICIT REAL*8(A-H,0-Z)
148 c
149 REAL*8 DH0(10)
1 50 c
151 T2 = T*T
152 T3 = T*T*T
153 TQ2 = DSQRT(T)
154 TQ = DSQRT(TQ2)
1 155 TQ3 = TQ * TQ2
1156 TQ5 = T * TQ
1 157 TQ6 = T * TQ2
1 158 TQ7 = T * TQ3
1 159 c
1 160 DH0(1) = 104.18*T+0.23276*T2-51714.3
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161 C
162 DHO(2) = -672.B7*T+111.25*TQ5-0.449495*TQ7+6477.6*TQ2
163 -39442.6
1 64 C
165 DH0(3) = 39.06*T+0.102558D7/TQ2-0.10727D8/T+0.4102D9/T2
166 -39672.7
167 C
1 68 37.432*T+0.0080408D-3*T2*TQ2+0.35714D6/TQ2
169 -0.23688D7/T - 23906.6
170 DH0(4) =
171 -3.7357*T+2.0353*TQ6-2.0517D-2*T2
172 +0.008066D-4*T3 - 7556.3
173 DHO(5) =
174 143.05*T-46.432*TQ5+5.51667*TQ6-0.0184 94 5*T2
175 - 11876.4
176 DHO(6) •
1 77 56.505*T-0.8889D5*TQ+0.1165D6*DLOG(T)
178 + 0.11214D7/TQ2 - 376187.0
179 DH0(7) =
180 81.546*T-15.0144*T5+0.313137*TQ7-0.0213 3*T2
181 -10509.4
182 DHO(8) -
183 69.145*T-0.0127328*TQ7-0.40154D4*TQ2
1 84 +0.223586D5*TQ - 43912.9
185 DHO(9) =
186 59.283*T-0.11397 3*T6-0.141226D4*TQ2
187 - 0.149778D6/TQ2 + 15975.8
188 C DHO(10)=
189 C
1 90 RETURN
191 END