Knock Mitigation on Boosted Controlled Auto-Ignition Engines with
Fuel Stratification and Exhaust Gas Recycling
by
Wen Sang
B.S., Automotive Engineering
Nanjing University of Aeronautics and Astronautics, 2008
M.S., Power Machinery Engineering
Tongji University, 2010
Submitted to the Department of Mechanical Engineering
in partial fulfillment of the requirements for the degree of
Doctor of Philosophy in Mechanical Engineering
at the
AF C$X\E,
MASSACHUSETS 1W?7ME,
OF TECHNOLOGY
MAY0E8 201
MASSACHUSETTS INSTITUTE OF TECHNOLOGY
LIBRARIES
February 2014
@ 2013 Massachusetts Institute of Technology
All rights reserved
Signature of Author ...............
Department of Mechanical Engineering
September 1 st 2013
---
Certified by ......................................................
--...........................................................................
Wai K. Cheng
Professor of Mechanical Engineering
Thesis Adviser
-...........-.....-......-..-.-..-----.--.--
Accepted by .................................
David E. Hardt
Students
on
Graduate
Committee
Department
Chairman,
1
Knock Mitigation on Boosted Controlled Auto-Ignition Engines with
Fuel Stratification and Exhaust Gas Recycling
by
Wen Sang
Submitted to the Department of Mechanical Engineering on September 1 s, 2013 in partial
fulfillment of the requirements for the degree of Doctor of Philosophy in Mechanical Engineering
Abstract
This research is carried out to understand the mechanism of using fuel stratification and Exhaust
Gas Recycling (EGR) for knock mitigation on boosted Controlled Auto-Ignition (CAl) engines. Experiments
were first conducted on Rapid Compression Machine (RCM) to profile the ignition characteristic of the
specific fuel used, and to explain the dilution effects of air and inert gas. Then the effect of fuel
stratification and EGR were systematically examined on a production engine (modified 1.9 L Renault F9Q
B800 common rail diesel engine) based test bench. The engine performance was interpreted with the
auto-ignition fundamentals to sort out the intrinsic links among CAI engine knock propensity, engine
operational parameters, and fuel stratification as well as EGR dilution extent. The nature of CAI engine
knock, the metric of the phenomenon, and the theoretical rationales behind using fuel stratification and
EGR for heat release control are reviewed before the experiment results are reported.
RCM tests show that the sensitivity of fuel ignition delay to equivalence ratio varies with the
ignition temperature, and higher sensitivity in the NTC region is preferred to make fuel stratification
useful. With fixed fuel concentration, air dilution slightly reduces the ignition delay, while inert gas
dilution could increase the ignition delay by a factor of 5. Inert gas dilution was found slowing down the
fast heat release effectively for ignition temperature around NTC region. This indicates strong effect of
EGR for CAI combustion knock mitigation.
Engine tests demonstrates that fuel stratification has high potential for CAI knock mitigation, but its
effect heavily depends on the extent of fuel stratification, engine configuration, and in-cylinder
conditions. While 80% improvement on knock performance can be achieved with mid-compression
stroke direct injection (DI), 400% higher knock intensity could also occur for late Dl. EGR was found
effective in retarding combustion phasing and reducing knock intensity, attribute to its effect on both incylinder temperature control and heat release curbing, yet misfire could happen with too much EGR.
With dual injections, the ratio of premixed fuel to directly injected fuel decreases the effect of fuel
stratification in all aspects. Higher intake temperature deteriorates the knock performance. Higher
engine speed retards the combustion phasing and enhances the fuel stratification extent and effect.
Analysis shows that CAI knock tendency is largely determined by the in-cylinder temperature
governed by combustion phasing, and many factors directly or indirectly influences the results. The
primary effect of fuel stratification is on combustion phasing, although the heat release rate is also
affected at the same combustion phasing. To better take advantage of fuel stratification and EGR for CAI
knock mitigation, the engine operating parameters have to be in the right range. This research work
could serve as a reference for future development of CAI engines with capability of knock free high load
operations.
Thesis Advisor: Wai K. Cheng
Title: Professor of Mechanical Engineering
3
(this page intentionally left blank)
4
Acknowledgments
My three years at MIT was an experience of dream coming true. Being able to work on cutting edge
research topic in field, constantly backed up by all the excellent resources and people, and learning so
much great knowledge and hands on skills everyday, I couldn't imagine a better academic life. None of
these are possible without the tremendous support from my advisor Prof. Wai K. Cheng, past and
current members of the MIT community, my friends, and my family.
Prof. Wai K. Cheng gave me the very opportunity to come to MIT, and this changed the trajectory of
my life without a doubt. No words can describe my appreciation to you. Your patient tutoring and
generous funding empowered me throughout my study at MIT, and you are always there to help me out
no matter it's simple problem with the electronic circuit of the experimental apparatus or hard sciences
that no one has ever figured out. Your firm support in the critical moments encouraged me to move on
to progress rather than lose myself in hesitation or confusion. You taught me things in the profession
and beyond it, as how to succeed in future as a person. Thank you Professor Cheng!
My thesis committee members Prof. William H. Green and Prof. John G. Brisson guided and
supported my work extensively. Thank Prof. Green for lots of excellent ideas and guidance on my work,
and generous support for letting me taking advantage of your research resources. Your great knowledge
in engine and combustion helped me look deeper into my work. Thank Prof. Brisson for your mentoring
on my thermal science fundamentals and fantastic questions from your great insight into my work. I
enjoyed the discussions with you and learned a lot from you all the time.
Past and current members of MIT Sloan Automotive Lab made my work fruitful and enjoyable.
Thank Prof. John Heywood for the great discussions on and after lab seminars that drove me progress.
Thank Dr. Tian Tian for all the great support and advices on life at and after MIT. Thank Dr. Amir Maria,
who laid the foundation of all my experiment systems, taught me almost hand by hand for so many
things, and set the example of an excellent MIT engineering PhD for me. Thank Raymond Phan, Thane
DeWitt for the great support on my experiments, and Janet Maslow for the great support on lab
administrative matters. Thank all the Sloan Lab mates for your generous help and friendship. Your great
work and personalities always inspire and encourage me.
The MIT community is so great and unique that I've benefited a lot from it in too many ways. Thanks
to all the good people in the community, my fantastic friends, and schoolmates. You enriched my life
and shaped my future on a fundamental level.
Special thanks to Wenlan Tian, my girl friend, who has always been there sharing my joy and support
me to get through the tough times.
My parents, Fengjie Sang and Xiaoqin Bai, thank you for your selfless love to me. Your support on all
my life decisions backed me to explore the world and experience MIT bravely. All my past and future
accomplishments are to make you proud.
5
(this page intentionally left blank)
6
Table of Contents
ABSTRACT ..............................................................................................................................................................
3
ACKNOW LEDGM ENTS ............................................................................................................................................
5
TABLE OF CONTENTS ..............................................................................................................................................
7
LIST OF FIGURES ...................................................................................................................................................
11
LIST OF TABLES ....................................................................................................................................................
17
NOM ENCLATURE .................................................................................................................................................
19
CHAPTER 1.
21
INTRODUCTION ..........................................................................................................................
1 .1 C A I E N G IN E K N O C K ...............................................................................................................................................
23
1.2 EFFORTS ON KNOCK MITIGATION .............................................................................................................................
25
1.2.1
Intake Boost ...........................................................................................................................................
25
1.2.2
Exhaust Gas Recycle ...............................................................................................................................
27
1.2.3
Therm al Stratification............................................................................................................................
28
1.2.4
Fuel Stratification...................................................................................................................................
32
1.2.5
Other Strategies .....................................................................................................................................
34
1.3 RESEARCH M OTIVATION.........................................................................................................................................35
1.4 RESEARCH APPROACH ............................................................................................................................................
36
1.4.1
Step I
Fuel Characteristics Investigation..........................................................................................
36
1.4.2
Step II- Engine Tests with Fuel Stratification and EGR.....................................................................
36
1.4.3
Step III - Engine Performance Interpretation.....................................................................................
37
-
OBJECTIVES ...........................................................................................................................................
37
1 .6 R EPO RT O U TLINE ..................................................................................................................................................
37
1.5 RESEA RCH
CHAPTER 2.
EXPERIM ENT APPARATUS ..........................................................................................................
39
2.1 RAPID COMPRESSION M ACHINE...............................................................................................................................39
2.1.1
RCM Specifications.................................................................................................................................
2.1.2
M IT Vertical RCM Working Procedure.................................................................................................41
2.1.3
Test Condition Set-up and Data Acquisition .......................................................................................
2.2 ENGINE EXPERIMENT SYSTEM .................................................................................................................................
2.2.1
Engine Specifications..............................................................................................................................47
7
40
44
46
2.2.2
Intake and Exhaust System ....................................................................................................................
48
2.2.3
Fuel Delivery System ...............................................................................................................................
51
2.2.4
Engine Controls and Data Acquisition ................................................................................................
54
CHAPTER 3.
NATURE OF CAI ENGINE KNOCK AND ITS M ITIGATION ...............................................................
57
3.1 THE PHYSICS OF CAI ENGINE KNOCK .........................................................................................................................
57
3.1.1
Therm odynam ics of In-cylinder Pressure Oscillation..........................................................................
3.1.2
Acoustics of the CAI Engine Knock..............................................................
.......
......... 58
3.1.3
The Energy Associated with CAI Engine Knock ...................................................................................
62
.........
57
3.2 CAI ENGINE KNOCK M ETRICS...................................................................................................................................63
3.2.1
Selecting the Proper Knock M etric .....................................................................................................
63
3.2.2
Relating the Acoustics M etric with Com bustion M easurem ent ..........................................................
65
3.3 CONTROLLING THE CAI ENGINE KNOCK .....................................................................................................................
67
3.3.1
The Necessity of Extending Burn Duration ..........................................................................................
67
3.3.2
The Potential of Stratification and EGR ...............................................................................................
70
3.3.3
Effectively Coordinating the Strategies ..............................................................................................
73
3.4 SUMMARY ...........................................................................................................................................................
CHAPTER 4.
75
FUEL IGNITION CHARACTERISTICS AND DILUTION EFFECTS .....................................................
77
4.1 FUEL IGNITION BEHAVIOR AND ITS M EASUREMENT...................................................................................................
77
4.1.1
Physics of Fuel Ignition Behavior............................................................................................................77
4.1.2
M easurem ent of Ignition Delay and M PRR ........................................................................................
4.1.3
Ignition Delay Curves..............................................................................................................................79
78
4.2 IGNITION W ITH FIXED CHARGE DENSITY.....................................................................................................................82
4.3 IGNITION W ITH FIXED FUEL DENSITY .........................................................................................................................
4.3.1
N2 as Diluent...........................................................................................................................................85
4.3.2
Sim ulated Com bustion Products as Diluent .......................................................................................
4.4 SUMMARY .........................................................................................................................................................
CHAPTER 5.
BOOSTED CAI ENGINE WITH FUEL STRATIFICATION AND EGR ..................................................
5.1 RESPONSES TO STRATIFICATION AT COMPRESSION RATIO = 19 ...............................................................................
85
96
104
107
107
5.1.1
M ap of Operation Range for the CR19 Engine .....................................................................................
107
5.1.2
Single Pulse Direct Injection .................................................................................................................
110
8
5.1.3
PartialFuel Stratification.....................................................................................................................
5.2 RESPONSES TO STRATIFICATION AT COMPRESSION RATIO = 15 ....................................................................................
117
126
5.2.1
Working M ap of the CR15 Engine ........................................................................................................
127
5.2.2
Low CR P FS Base Case ..........................................................................................................................
128
5.2.3
Fuel Split Ratio
Impact .........................................................................................................................
135
5.2.4
Intake Tem perature Impact .................................................................................................................
141
5.2.5
Engine Speed Im pact ............................................................................................................................
149
5.3 CONCLUSION......................................................................................................................................................153
CHAPTER 6.
EFFECT OF FUEL STRATIFICATION ON CAl COM BUSTION ..........................................................
6.1 STRATIFICATION AND ITS IMPACT TO CA10 ........................................................................................
155
155
6.1.1
The Existence of Fuel Stratification with DI S01....................................................................................155
6.1.2
A conceptual M odel of Fuel Stratification on CA10..............................................................................158
6.2 CHARGE TEMPERATURE, STRATIFICATION, AND HEAT RELEASE.....................................................................................162
6.3
6.2.1
Com bustion Phasing, Charge Tem perature, and Heat Release............................................................163
6.2.2
Stratification and Burn Duration ..........................................................................................................
164
SENSITIVITY OF IGNITION TO FUEL STRATIFICATION ....................................................................................................
167
6.3.1
Ignition Delay Data ..............................................................................................................................
167
6.3.2
Com pression Process in the Engine Cycle .............................................................................................
169
6.4 CONCLUSION......................................................................................................................................................174
CHAPTER 7.
SUM M ARY AND CONCLUSION..................................................................................................
175
7.1 CAl ENGINE KNOCK AND POTENTIAL SOLUTIONS.......................................................................................................175
7.2 FUEL IGNITION CHARACTERISTICS ...........................................................................................................................
176
7.3 ENGINE PERFORMANCE W ITH FUEL STRATIFICATION AND EGR ....................................................................................
177
7.4 THE EFFECT OF FUEL STRATIFICATION AND EGR ........................................................................................................
178
7.5 CONCLUSION......................................................................................................................................................180
APPENDIX: SPECIFICATIONS OF HALTERM ANN 437 GASOLINE ..........................................................................
181
REFERENCE ........................................................................................................................................................
183
9
(this page intentionally left blank)
10
List of Figures
23
Figure 1-1
Pressure and microphone signals (recorded simultaneously) for a knocking CAI cycle [4] ...............
Figure 1-2
Impact of intake pressure on the IMEP in a CAI engine using excess air (bottom line) and cooled EGR
26
(top line); reprinted from Dec [20].....................................................................................................
Figure 1-3
Knock-limited NIMEP versus CAMPRR; fuel comparison; port injection, MAP = 1.9 bar, 1900 rpm;
0
re p rinte d fro m A m ir [1]..........................................................................................................................3
Figure 1-4
Ignition delay curve corresponding to (a) Fuel #1, (b) Fuel #2, and (c) Fuel #3; MAP = 1.9 bar,
CAM PRR 357 CAD; reprinted from M aria [1] .........................................................................................
31
Figure 1-5
Equivalence ratio sensitivity of gasoline and PRF73 under different intake pressures; reprinted from
33
De c [4 1 ] ..................................................................................................................................................
Fig u re 2 -1
M IT ve rtica l R C M ....................................................................................................................................
40
Figure 2-2
M ixture preparation unit (M PU) schem atic [46].................................................................................
41
Figu re 2-3
RC M co re part [4 6].................................................................................................................................4
3
Figure 2-4
RCM hydraulic netw ork schem atic [46] ..............................................................................................
44
Figure 2-5
Pressure traces from the RCM tests...................................................................................................
45
Figure 2-6
Engine experim ent system .....................................................................................................................
46
Figure 2-7
Renault F9Q B800 common rail diesel engine ..................................................................................
47
Figure 2-8
M odified intake and exhaust system ................................................................................................
48
Figure 2-9
Engine intake system schem atic .......................................................................................................
49
Figure 2-10
Fuel delivery distribution for varied DI SO I .......................................................................................
52
Figure 2-11
PFI
injecto r calibration chart ..................................................................................................................
53
Figure 2-12
Pressure supply and fuel return of the PFI injection system...............................................................53
Figure 2-13
Pressure traces and pressure oscillation under knock condition [49] ..............................................
Figure 3-1
Frequencies of the in-cylinder gas drum modes of oscillation; reprinted from Andreae [4].............60
Figure 3-2
Power spectrum density distributions for first circumferential mode (4-8 kHz) and higher order modes
60
(8-25kHz) for SI and HCCI operation; reprinted from Eng [28] .........................................................
Figure 3-3
Spectral powers from pressure transducer and microphone signals of the four subjectively classified
knock cases. The 1st order mode is for the 4-8 KHz range; the higher order mode is for the 8-20 KHz
...................... 61
range; Reprinted from A ndreae [4]...........................................................................
Figure 3-4
Power spectral density of sound radiation form the engine excited by an impulse; reprinted from
. ............... ........... 6 2
A n d re a e [4 ] ............................................................................................................
Figure 3-5
Microphone signal amplitude versus knock intensity for a large data set on log-log scale; reprinted
............................. . -- . ...... ........... 65
fro m M a ria [4 9 ].......................................................................
11
55
Figure 3-6
Correlation plot of MPRR to engine combustion parameters; reprinted from Wildman [51]...........69
Figure 3-7
Schematic showing the ignition controlled heat release process [52] .............................................
71
Figure 4-1
Sample pressure trace versus time from RCM tests .........................................................................
79
Figure 4-2
Sample ignition delay curve; reprinted from Maria [1].....................................................................
80
Figure 4-3
Chain reactions and NTC region mechanism; reprinted from Hahn [54]..........................................
82
Figure 4-4
Ignition delay with fixed charge density 0.75 Kmol/m3 ...................................
Figure 4-5
Ignition delay of iso-octane in high temperature regime; reprinted from K. Fieweger [55] .............
84
Figure 4-6
Ignition delay----air dilution effect with inert gas dilution rate=0 ....................................................
87
Figure 4-7
Ignition delay----air dilution effect with inert gas dilution rate=0.2 ..................................................
87
Figure 4-8
Ignition delay----air dilution effect with inert gas dilution rate=0.4 ..................................................
88
Figure 4-9
MPRR---- air dilution effect with inert gas dilution rate=0.................................................................90
Figure 4-10
MPRR---- air dilution effect with inert gas dilution rate=0.2..............................................................90
Figure 4-11
MPRR---- air dilution effect with inert gas dilution rate=0.4..............................................................91
Figure 4-12
Ignition Delay---- inert gas dilution effect with equivalence ratio=1.0 ..............................................
92
Figure 4-13
Ignition Delay ---- inert gas dilution effect with equivalence ratio=0.8 ............................................
92
Figure 4-14
Ignition Delay ---- inert gas dilution effect with equivalence ratio=1.0 ............................................
93
Figure 4-15
MPR R---- inert gas dilution effect with equivalence ratio=1.0..........................................................
93
Figure 4-16
MPRR---- inert gas dilution effect with equivalence ratio=0.8..........................................................
94
Figure 4-17
MPRR----boost effect with equivalence ratio=0.6..............................................................................94
Figure 4-18
A ir dilution vs. inert gas dilution. ...........................................................................................................
95
Figure 4-19
Dilution rate dependence of inert gas dilution effectiveness ............................................................
96
Figure 4-20
Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=1.0 ................................................
97
Figure 4-21
Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=0.8 ................................................
98
Figure 4-22
Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=0.6 ................................................
98
Figure 4-23
MPRR-Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=1.0 .....................................
99
Figure 4-24
MPRR-Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=0.8 .....................................
99
Figure 4-25
MPRR-Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=0.6 ...................................
100
Figure 4-26
Ignition Delay----Fuel density Impact with equivalence ratio=1.0 .......................................................
101
Figure 4-27
Ignition Delay----Fuel density Impact with equivalence ratio=0.8 .......................................................
101
Figure 4-28
Ignition Delay----Fuel density Impact with equivalence ratio=0.6 ..................................................
102
12
. .. .. . .. .. . .. .. . .. .. .. .. .. . .. . . . .
83
Figure 4-29
M PRR----Fuel density Im pact with equivalence ratio=1.0....................................................................102
Figure 4-30
M PRR----Fuel density Im pact with equivalence ratio=0.8....................................................................103
Figure 4-31
M PRR----Fuel density Im pact with equivalence ratio=0.6....................................................................103
Figure 5-1
M ap of the test engine w orking range .................................................................................................
108
Figure 5-2
KI variation w ith SOI tim ing for high CR single pulse DI tests ..............................................................
111
Figure 5-3
KI variation with SOI timing for high CR single pulse DI tests (log scale) .............................................
111
Figure 5-4
M PRR variation with SOI tim ing for single pulse DI tests.....................................................................112
Figure 5-5
CA10 and CA 50 variation with SOI for single pulse DI tests ................................................................
113
Figure 5-6
Burn duration variation w ith SOI for single pulse DI tests ...................................................................
114
Figure 5-7
NOx em ission variation with SOI for single pulse DI tests ....................................................................
115
Figure 5-8
Pressure traces for CR19, EGR=40%, Single Pulse DI tests ...................................................................
116
Figure 5-9
KI variation w ith DI SO Itim ing for PFS tests.........................................................................................118
Figure 5-10
KI variation w ith DI SOI tim ing for PFS tests (log scale) .......................................................................
119
Figure 5-11
M PRR variation w ith DI SOI tim ing for PFS tests ..................................................................................
120
Figure 5-12
CA10 and CA50 for CR19, EGR=40%, PFI and DI dual injection tests ...................................................
121
Figure 5-13
Burn duration for CR19 PFS tests .........................................................................................................
122
Figure 5-14
NOx em ission variation with DI SOI timing for CR19 PFS tests.............................................................122
Figure 5-15
GIM EP variation with DI SOl tim ing for CR19 PFS tests........................................................................123
Figure 5-16
Pressure traces for CR19, EGR=40%, PFI and DI dual injection tests I .................................................
123
Figure 5-17
Pressure traces for CR19, EGR=40%, PFI and DI dual injection tests II ................................................
124
Figure 5-18
Pressure traces for CR19, EGR=40%, PFI and DI dual injection tests III ...............................................
124
Figure 5-19
Heat release process comparison for CR19, EGR=40%, PFI and Dl dual injection tests.......................125
Figure 5-20
Working range of the low CR engine (RPM1200, intake T=80 "C, P=1.8 bar) ......................................
Figure 5-21
KI of the base case on low CR engine (RPM1200, intake T=80 *C, P=1.8 bar)......................................129
Figure 5-22
MPRR of the base case on low CR engine (RPM1200, intake T=80 "C, P=1.8 bar)...............................130
Figure 5-23
CA10 and CA50 of CR15 base case tests ..............................................................................................
131
Figure 5-24
Burn Duration from CR15, PFI and DI dual injection tests ...................................................................
132
Figure 5-25
Pressure traces from CR15, EGR=15%, PFI and DI dual injection tests I ..............................................
132
Figure 5-26
Pressure traces from CR15, EGR=15%, PFI and DI dual injection tests
11 .............................................
133
Figure 5-27
Pressure traces from CR15, EGR=15%, PFI and DI dual injection tests Ill ............................................
133
13
128
Figure 5-28
Heat release process comparison for CR15, EGR=15%, base case tests ..............................................
Figure 5-29
GIMEP comparison for CR15, EGR=15%, base case tests.....................................................................135
Figure 5-30
GIM EP for PFI/DI ratio of 40/60 and 80/20..........................................................................................136
Figure 5-31
KI for PFI/DI ratio study w ith EGR=0 and 15% ......................................................................................
137
Figure 5-32
M PRR for PFI/DI ratio study with EGR=0 and 15% ...............................................................................
138
Figure 5-33
CA10 for PFI/DI ratio study w ith EGR=0 and 15% ................................................................................
139
Figure 5-34
CA50 for PFI/DI ratio study w ith EGR=0 and 15% ................................................................................
139
Figure 5-35
Burn Duration for PFI/DI ratio study with EGR=0 and 15%..................................................................140
Figure 5-36
GIMEP for intake temperature impact study, EGR=0 and 10%............................................................142
Figure 5-37
KI for intake temperature impact study, EGR=0 and 10%....................................................................143
Figure 5-38
MPRR for intake temperature impact study, EGR=0 and 10%.............................................................144
Figure 5-39
CA10 for intake temperature impact study, EGR=0 and 10% ..............................................................
145
Figure 5-40
CA50 for intake temperature impact study, EGR=0 and 10% ..............................................................
145
Figure 5-41
Burn Duration for intake temperature impact study, EGR=0 and 10%................................................146
Figure 5-42
Excessive air coefficient comparison for tests with different intake temperatures ............................
Figure 5-43
KI of tests with intake temperature 100 "C, CR 15, RPM 1200, MAP 1.8 bar.......................................147
Figure 5-44
MPRR of tests with intake temperature 100 *C, CR 15, RPM 1200, MAP 1.8 bar................................148
Figure 5-45
Com bustion Phase w ith intake tem perature 100 "C............................................................................148
Figure 5-46
Burn duration w ith intake tem perature 100 C ...................................................................................
149
Figure 5-47
KI of tests for engine speed im pact study ............................................................................................
150
Figure 5-48
M PRR of tests for engine speed im pact study .....................................................................................
151
Figure 5-49
Combustion phase of tests for engine speed impact study .................................................................
151
Figure 5-50
Combustion phase of tests for engine speed impact study .................................................................
152
Figure 6-1
HC and NO emissions with DI SOI variation for tests with PFI/DI=30/70, Tin = 80 *C, Pin = 1.8 bar,
engine speed 1200 rpm, and total fuel 23 mg/cycle on CR19 engine..................................................157
Figure 6-2
HC and NO emissions with DI SOl variation for tests with PFI/DI=40/60, Tin = 80 *C, Pin = 1.8 bar,
engine speed 1200 rpm, and total fuel 23 mg/cycle on CR 15 engine .................................................
134
146
157
Figure 6-3
CA10-DI SO I for single pulse DI test in Chapter 5.................................................................................160
Figure 6-4
CA10, compression temperature and pressure with DI SOl for CR15 base case test ..........................
161
Figure 6-5
MPRR vs. CA10 and Temperature at CA10 for CR15 base case tests ...................................................
163
Figure 6-6
MPRR of mixture ignited at CA10 temperatures of CR15 base case tests ...........................................
164
14
Figure 6-7
Burn Duration vs. CA10 of CR15 base case tests..................................................................................165
Figure 6-8
Burn Duration vs. CA10 of CR15 base case tests zoom -in....................................................................165
Figure 6-9
MPRR vs. CA10 for stratification comparing test set of CR 15 base case with lower PFI/DI ratio at
4 0 / 6 0 ....................................................................................................................................................
1 66
Figure 6-10
CHEM KIN M odel validation w ith experim ents.....................................................................................168
Figure 6-11
CH EM KIN M odel results for P=0.25 ....................................................................................................
168
Figure 6-12
CH EM KIN M odel results for D=0.4 ......................................................................................................
169
Figure 6-13
Pressure trace comparison for simulation and experiment for test with CR19, PFI fuel=16 mg/cycle,
170
Tin= 80 *C, Pin=1.9 bar, engine speed 1200 rpm , no EGIR .....................................................................
Figure 6-14
Pressure trace comparison for simulation and experiment for test with CR15, PFI fuel=23 mg/cycle,
170
Tin= 80 'C, Pi,=1.8 bar, engine speed 1200 rpm , no EGR .....................................................................
Figure 6-15
In-cylinder conditions along compression process for CR 19 engine ...................................................
171
Figure 6-16
In-cylinder conditions along compression process for CR 15 engine ...................................................
172
Figure 6-17
Ignition delay curve for CR 15 base case test with 15% EGR ...............................................................
173
Figure 7-1
Links between CAl engine knock performance and operation conditions...........................................179
15
(this page intentionally left blank)
16
List of Tables
Table 1-1
High load range of CAl engines from literature ..................................................................................
24
Table 1-2
Typical BM EP values of conventional IC engines ................................................................................
24
Table 1-3
EGR effects on auto-ignition for different fuels.................................................................................
28
Table 2-1
RCM design specifications and component limitations [46]...............................................................41
Table 2-2
Engine specifications [47] .......................................................................................................................
48
Table 2-3
Engine control variable sum m ary .......................................................................................................
54
Table 3-1
Kncok criteria in previous research.....................................................................................................
64
Table 4-1
RCM Run M atrix----N 2 as Diluent ....................................................
17
...........
...........
............ 86
(this page intentionally left blank)
18
Nomenclature
ACRONYMS
ABDC
BDC
BMEP
BTDC
CA
CAI
CR
DI
EGR
GIMEP
HCCI
HLL
HTHR
IC
IMEP
KI
LHV
LTHR
MAP
MEP
MPRR
MPU
NIMEP
NTC
NVH
NVO
PCCI
PFI
PFS
RCM
RPM
RI
SOI
TDC
After Bottom Dead Center
Bottom Dead Center
Brake Mean Effective Pressure
Before Top Dead Center
Crank Angle
Controlled Auto-Ignition
Compression Ratio
Direct Injection
Exhaust Gas Recirculation
Gross Indicate Mean Effective Pressure
Homogeneous Charge Compression Ignition
High Load Limit
High Temperature Heat Release
Internal Combustion
Indicated Mean Effective Pressure
Knock Intensity
Lower Heating Value
Low Temperature Heat Release
Manifold Air Pressure
Mean Effective Pressure
Maximum Pressure Rise Rate
Mixture Preparation Unit
Net Indicated Mean Effective Pressure
Negative Temperature Coefficient
Noise Vibration and Harshness
Negative Valve Overlap
Premixed Charge Compression Ignition
Port Fuel Injection
Partial Fuel Stratification
Rapid Compression Machine
Revolutions Per Minute
Ringing Index
Start of Injection
Top Dead Center
19
SYMBOLS
[A]
B
-r
D / Phi
f
y
C
Pf
CA10
CA50
CAMPRR
CP
Cv
n
EA
P
Pin
AP
R
t
T
Tin
V
Vd
Concentration of species A
Engine Bore
Excessive air coefficient
Ignition delay time
Fuel-air equivalence ratio
Frequency
Specific heats ratio
Speed of sound
Fuel concentration
Crank angle at 10% chemical heat release
Crank angle at 50% chemical heat release
Crank angle at the location of MPRR
Constant pressure specific heat
Constant volume specific heat
Molar concentration
Liven-good Wu integration; or pressure oscillation energy flux
Activation Energy
Pressure
Intake pressure
Pressure oscillation
Ideal Gas Constant
Time
Temperature
Intake temperature
Volume
Displaced cylinder volume
Unless otherwise stated, all engine crank angle degree values given in this report are referenced as After
Bottom Dead Center intake.
20
Chapter 1. Introduction
Controlled
Auto-Ignition
(CAI)
combustion,
sometimes
referred
as
Homogeneous
Charge
Compression Ignition (HCCI), is a novel combustion mode in internal combustion (IC) engines in contrast
to spark ignition combustion for gasoline and compression ignition combustion for diesel. It is different
from either while it has certain features of both: 1) the combustible charge is mostly premixed, yet
sometimes the fuel is injected directly into the cylinder; 2) typical in-cylinder mixture is globally lean; 3)
the chemical kinetics driven ignition occurs automatically upon compression rather than initiated by
spark or fuel injection; 4) no flame propagation is involved in the heat release process, therefore the
combustion temperature can be significantly lower, and the duration of combustion is not governed by
the speed of turbulent or diffusion flame; 5) the compression ratio can be low or high depending on the
fuel and the engine; 6) power output is controlled by the amount of fuel put into engine cylinder rather
than air throttling as it's typically lean combustion [1].
Since the phenomenon first observed in late 1970s [2], lots of interest has been drawn from
researchers all over the world [3]. This is because the operation of CAI engine has a great potential for
clean and high efficiency combustion for a variety of fuels:
1.
NOx emissions are much less in CAI engines than conventional engines; the reduction is
attributed to much lower combustion temperature achieved from the lean mixture and
flameless combustion.
2.
Particulate emissions of CAI engines are much lower than Diesel engines with the well-mixed incylinder charge burning without fuel-rich diffusion flame.
3.
Pumping loss is lower on gasoline CAI engines than SI engines for better mechanical efficiency,
as power output control is realized without throttling under high dilution and lean burn
conditions.
4.
Thermodynamic efficiency can be improved by having lean mixture with higher specific heat
ratio, which is an effect due mostly to the lower combustion temperature.
5.
Compression Ratio (CR) can be optimized for better thermodynamic efficiency (to suppress
knock, SI engine normally has a lower CR than ideal; Diesel engine normally has a higher CR than
ideal to enhance the cold-start.).
21
6.
Fuels with very different reactivity are all found capable of CAI operation with corresponding
engine design, and this could give much larger space for the alternative fuels to impact.
However, the industrialization of CAI engine is still bottlenecked by a few technology challenges:
1.
Ignition timing control. Ignition timing has significant impact on engine performance in thermal
efficiency, operation stability, NVH, and emissions. Comparing with the conveniently controlled
ignition timing of SI engines by spark and diesel engines by fuel injection, CAI engines are in a
much critical situation.
The time required for the mixture to ignite in CAI engines heavily
depends on chemical kinetics, which can be impacted by fuel properties, intake temperature &
pressure, exhaust gas recycle (EGR) rate, valve timing, non-uniformities in the mixture, and lots
of other factors. To have a reliable CAI combustion system for a wide range working conditions,
all of these variables need to be coherently adjusted. This requires much more advanced
understanding of the in-cylinder going on and control techniques.
2.
HC and CO emissions. With mostly premixed charge but low burned gas temperature, fuel
stored in crevices during the compression stroke couldn't be well consumed when they reenters the cylinder during expansion. This causes significant HC and CO emissions on CAI
engines. CO to CO 2 conversion is also worse in low load operations due to the low peak
temperatures. This could be compounded with reduced after-treatment efficiency due to low
exhaust temperature.
3.
Limited operation range. Comparing with conventional engines, CAI engines typically have more
critical low load operation requirements, as the temperature needed for auto-ignition is much
elevated than what's available on engine without intake heating. When hot exhaust gas is
recycled/captured for strengthening the low load stability, the dilution effect on combustion
phasing retarding kicks in as an extra concern. In high load regime, potential pressure
oscillations caused by abrupt pressure rise due to high fueling rate starts constraining the
operation of CAI combustion. When the amount of fuel inside engine cylinder becomes
sufficiently large that equivalence ratio approaches to stoichiometric value, CAI combustion also
begins to lose its the advantage on low NOx emissions.
Professionals and scientists all over the world have spent tremendous effort to tackle above
problems [3]. The author has been focusing on exploring the potential of fuel stratification and EGR for
heat release rate control and High Load Limit (HLL) expansion on light duty gasoline CAI engines under
22
boosted conditions. In this chapter, the research motivation, objectives, and approaches are to be
clarified after a brief review of the topic and previous work. The structure of this report will be outlined
in the end of the chapter.
1.1
CAI Engine Knock
By the nature of CAI combustion, the flameless heat release process happens in a time period much
shorter than the conventional IC engines. This is because unlike on conventional engines, where the
combustion completion depends on the speed of flame propagation governed by the heat and mass
transfer processes, CAI combustion occurs volumetrically across the engine cylinder, and the completion
of combustion is limited only by the rate of chemical reactions under well mixed conditions. This leads
to a vital problem that when more fuel is injected into the cylinder for higher power output, the fast
heat release could result in drastic pressure rise and therefore pressure waves inside engine cylinders.
The pressure waves will excite the engine structure to vibrate severely, which leads to unacceptable
noise radiation and potential engine damage. This is called CAI engine knock.
This issue bottlenecks the maximum amount of fuel that can be put into the cylinder, and thus the
HLL of CAI engines. To have an idea of the phenomenon, a pressure trace of CAI combustion under
knocking condition is shown in Figure 1-1.
50'*li
Mic-out
40
LI
30
.-
20
Pressure
10
U
1.124
1.126
1.128
1.13
1.132
1.1 34
Time (a
Figure 1-1
Pressure and microphone signals (recorded simultaneously) for a knocking CAI cycle [4]
23
The acoustic noise perceived simultaneously by a microphone above the firing cylinder is also shown
in the figure. The pressure oscillation and noise induced by the fast heat release are clearly seen [4]. The
physics of CAI engine knock and its measurement is reviewed in more details in Chapter 3.
Table 1-1 shows the high load range of CAI engines from recent years research literature. Easy to see
with many years of hard work on the topic [6~19], the high load limit of CAI engines are still very limited
comparing with conventional engines, as shown in Table 1-2 [5]. As mentioned above, engine knock is
one of the very constraints.
Table 1-1
High load range of CAI engines from literature
Research Group/Affiliation
Tsinghua University
U. Michigan
Sandia National Lab
U. Birmingham & Jaguar Car Ltd.
Tokyo City Univ.
MIT
Lund Univ. & Shell
Oak Ridge National Lab
MAHLE Powertrain
U. Michigan
GM
U. Wisconsin
AVL & Robert Bosch
Oak Ridge National Lab & Delphi
SAE Paper
2007-01-0195
2007-01-0204
2008-01-0054
2009-01-1101
2009-01-1419
2010-01-0162
2010-01-0607
2010-01-2172
2010-01-2196
2011-01-0888
2011-01-0899
2012-01-0376
2013-01-1655
2013-01-1665
MEP Upper Bound Reported (bar)
10
7
6.34
4
4.5
5
10
5
10
10
10
10
5
6.5
Some research work has demonstrated the possibility of CAI engines to hit the high IMEP values up
to 16 bar, yet very critical, or even unrealistic, requirements have to be satisfied regarding EGR, intake
temperature and pressure (e.g. higher than 3 bar), fuel type, fuel injection method, and many other
operation parameters [20~24]. This means fundamental research is urgently needed to better the
understanding of the effect of these techniques and to optimize the combustion control strategies on
CAI engines toward the direction of practical applications.
Table 1-2
Typical BMEP
Values
Max Torque
Max Rated Power
Typical BMEP values of conventional IC engines
SI Engines
Natural Aspiration
Turbocharged
8.5 -10.5 bar
7.5 "9.5 bar
12.5 ~17 bar
9 ~14 bar
24
Diesel Engines
Natural Aspiration
Turbocharged
7 ~9 bar
10 ~14 bar
- 7 bar
8.5 "~9.5 bar
Efforts on Knock Mitigation
1.2
To abate engine knock and expand the CAI combustion high load limit, various techniques and
methods have been investigated [6~24]. While some research focuses on reducing the heat release rate
by enhancing the inhomogeneity of the in-cylinder charge, some research tries for better knock
endurance with fixed pressure rise rate, and some other research attempts to slow down the chemical
reaction process by dilution or novel fueling strategies. This section gives a brief review of these works.
1.2.1
Intake Boost
Higher intake density offers opportunity for power output increase at the same level of lean
operation. However, it also render higher knock propensity: 1) more charge mass inside engine cylinders
gives greater pressure rise rate for the same combustion phase; 2) boost enhances the auto-ignition
process leading to higher pressure rise rate [20].
Early research by Christensen and Johansson in Lund Institute of Technology on a supercharged CAI
engine [25] showed significantly enhanced IMEP by intake boost, although corresponding adjustment of
intake heating is required to compensate for the advanced combustion phase under elevated manifold
air pressure (MAP). They also noticed that intake boost has different impact on different fuels. With
natural gas, the IMEP can be maximized up to 14 bar under absolute intake pressure of 3 bar, while for
iso-octane the maximum IMEP is 9.7 bar. The IMEP was further pushed up to 16 bar by using natural gas
as main fuel and iso-octane as pilot, combining with EGR and variable compression ratio [22].
With an turbocharged engine, Thomas Johansson et al. in Lund University demonstrated that
increased intake pressure could reduce the pressure rise induced combustion noise and lead to wider
operating range for CAI combustion [26]. However since limitation from engine ringing index, peak
pressure, NOx and soot emissions are not necessarily well aligned in all the speed-load points, further
work was found necessary to make intake boost practically effective for CAl HLL expansion.
John Dec and his colleague in Sandia National Lab found that increasing the intake pressure by
cooled EGR rather than air could greatly increase the attainable higher IMEP [20]. This is because by air
dilution, more available oxygen advances the combustion phase, while by using cooled EGR to displace
excess oxygen could slow down the auto-ignition process and therefore delay the combustion phase. At
an intake pressure of 3.25 bar, Dec achieved IMEP of 16.34 bar by increasing the EGR rate until unstable
combustion occurs, improving from the 5 bar IMEP with natural aspiration and the 6 bar IMEP at 3.25
25
bar MAP but no EGR dilution. This test is conducted by setting the knock limit as a maximum Ringing
Index of 5 MW/M 2 . The key of this accomplishment was attribute to the intermediate-temperature heat
release, which provided the ability of stable combustion with significantly retarded combustion phase
(19 "CA ATDC for the maximum IMEP point). A full map of the intake boost effect on CAI HLL expansion
with air and EGR as diluent is shown in Figure 1-2. Notice the intake temperature was varied for
combustion phase retarding to maximize the IMEP for air-diluted tests (bottom line) with intake
pressure lower than 180 kPa. With higher intake pressure, the intake temperature was fixed at 60 *C to
prevent condensation of water vapor in EGR gas, as well as to divert the combustion from potential lowtemperature heat release (LTHR) which requires large adjustments for maintaining combustion phase.
18
All Data Points
16
14
-*-Max
IMEPg, with EGR
v12
00
w 8
6
4
2
0 1
50
Figure 1-2
-u-Max IMEPg, no EGR
ing
1
100
5 MW/m
1
1
1
150
200
250
300
Intake Pressure [kPa abs.]
350
Impact of intake pressure on the IMEP in a CAI engine using excess air (bottom line) and
cooled EGR (top line); reprinted from Dec [20]
Amir Maria [1] from MIT also argued that intake pressure improves CAI engine knock performance
base on his own experiments and analysis on CAI knock theories proposed by Yelvington [27] and Eng
[28] respectively: 1) Higher MAP increases the chamber pressure and reduces the pressure oscillation
amplitude; 2) Higher chamber pressure that could be induced by boosted intake pressure also gives a
lower acoustics energy flux for the same pressure rise rate which extends the high load limit. The
reasoning behind these statements will be reviewed specifically in Chapter 3.
From above we see that the impact of intake boost to CAI combustion and its HLL is complex:
26
1.
Higher intake pressure provides the potential for higher fueling rate.
2.
Higher intake pressure advances the combustion phase and exposes CAI combustion under
higher risk of knock.
3.
Higher intake pressure could reduce the engine knock propensity from the acoustics
perspective.
Therefore further research work would
be needed for achieving more
intrinsic and
comprehensive understanding of the intake boost effect on CAI combustion HLL expansion.
1.2.2
Exhaust Gas Recycle
EGR has been extensively used for CAI combustion control and high load limit exploration. Internal
EGR created by negative valve timing provides hot residual to enhance the start-up and low load
operation capability [29, 30]. Cooled external EGR dilutes the fuel-air mixture and modifies the
properties of the in-cylinder charge to achieve higher load [20]. Sometimes the two types of EGR are
combined to reduce the engine knock propensity while maintaining stable combustion [31].
Sj6berg and Dec [32] explored the potential of EGR on controlling the LHTR of two-stage fuels, which
was found a promising fact to be exploited for CAI engine knock mitigation. They demonstrated that EGR
effectively counteracts auto-ignition and loses up the requirements on intake temperature control for
LHTR maintenance.
To thoroughly understand the thermodynamic and chemical effects of EGR on CAI combustion,
Sj6berg and Dec experimentally investigated the impact of real EGR, simulated EGR (complete
stoichiometric products (CSP) - N2 , C0 2 , and H20), dry EGR (CSP without H2 0) on the auto-ignition
process for gasoline and various PRF blends [33]. Several effects of EGR
are identified: 1)
thermodynamics cooling due to the modified charge specific heat by CO 2 and H20, retarding combustion
phase; 2) reduced concentration of oxygen as air replaced by EGR at given intake pressure, retarding
combustion phase; 3) presence of H20, enhancing auto-ignition; 4) presence of trace species,
suppressing auto-ignition. They found the responses of engine performance under different EGR types
and fuel types are not all the same. To identify the underlying mechanisms, experiments are also
conducted with individual EGR constituents (N2 , CO
2,
Table 1-3.
27
and H2 0). Their findings are summarized as in
Table 1-3
Effect Items
Thermodynamic cooling
[02] reduction
Presence of H20
Presence of Trace Species
EGR effects on auto-ignition for different fuels
One-Stage Fuels
(iso-octane and gasoline)
Effect on Auto-ignition Sensitivity
Retarding
High
Retarding
Low
Enhancement
Low
Enhancement
-
Two-stage Fuels
(PRF80 and PRF60)
Effect on Auto-ignition
Sensitivity
Retarding
Low
Retarding
High
Enhancement
High
Suppression
-
Wildman [34] from MIT systematically studied the impact of EGR and other factors such as intake
temperature, intake pressure, and valve timing on CAl HLL. The test engine was operated with
adjustable Negative Valve Overlap (NVO) via Variable Valve Actuation (VVA), and external EGR while
equivalence ratio was fixed at 1. He found using external EGR instead of NVO trapped residual gas for
charge dilution extended the HLL as the charge temperature was reduced. While intake boost and intake
heating raises the charge temperature, EGR offsets this effect. Under the constraint of a fixed maximum
pressure rise rate (MPRR), EGR combined with intake boost or intake heating wouldn't vary the HLL
effectively.
Concurrent research conducted by Scaringe [35] also found that external EGR has the effect of
retarding combustion phase by reducing charge temperature, but boosting the intake wouldn't
substantially help improve the CAl combustion high load limit with a fixed MPRR constraint. Amir Maria
pointed out that if Knock Intensity was set as the knock metric rather than MPRR, a larger difference
might be seen [1].
To summarize: EGR has the potential for heat release rate reduction, as fundamentally inert gas
dilution might extend the ignition delay and relax the heat release process. However, the multiple
effects of EGR on CAI combustion as shown above make adoption of the technology sophisticated.
Notice these effects can be of very different trend and sensitivity for different fuels. To properly exploit
the potential of EGR on CAI knock mitigation, fundamental studies of the dilution effects need to be
conducted. For specific combustion systems, EGR needs to be subtly implemented to benefit from
combustion phase retarding while maintaining stable combustion.
1.2.3
Thermal Stratification
28
Thermal stratification naturally exists in the engine cylinders due to the highly dynamic working
cycle, drastically changing in-cylinder temperature, and constant mass & heat transfer processes such as
gas exchange and fuel evaporation. Intentionally enhancing the temperature gradient among in-cylinder
mixture parcels could potentially increase the inhomogeneity of the ignition and heat release properties
of these parcels, thereby induce sequential ignition and extended burn duration for a lower pressure
rise rate. Many researchers have investigated the strategy of using thermal stratification for CAI
combustion control and HLL expansion.
Noda and Foster numerically studied the hydrogen fueled CAI combustion and found that by
introducing appropriate temperature inhomogeneity to the in-cylinder temperature, the combustion
duration can be extended and the heat release rate can be reduced [36].
Sj6berg [37] identified four potential causes of the naturally existed thermal stratification: "1) heat
transfer from the charge to the combustion chamber walls during compression, 2) the presence of hot
residuals pockets from the previous cycle as a result of incomplete mixing, 3) dynamic flow effects
during the induction stroke, and 4) vaporization of the fuel, especially if injected directly into the
cylinder." With the knowledge that wall temperature variation changes the heat transfer rate so that
the in-cylinder temperature gradient, he experimentally examined the impact of in-cylinder temperature
stratification by reducing the coolant temperature from 100 'C to 50 "C in a modified single cylinder CAI
engine. A drop of the Knock Intensity from 6.23 MW/m
2
to 3.71 MW/M
2
was achieved under constant
loading and a constant CA10. Noticed heat-transfer rates could also be increased by strengthen the air
swirl and the associated low velocities, he increased the intake swirl ration from 0.9 to 3.6 to enhance
the thermal stratification and further reduced the Knock Intensity to 2.95 MW/M2 . However, substantial
drop of IMEP was associated as a penalty from the increased heat loss.
Kakuho et al. [38] created thermal stratification in the CAI combustion mixture prior to ignition with
two separately heated intake ports by feeding the cylinder differently heated fresh charge. The
temperature distribution was visualized and measured with a laser induced fluorescence system. It was
found that the stratified charge case had an earlier start of combustion but longer burn duration than
the homogeneous case.
Kuboyama et al. [39] created in-cylinder thermal stratification on a "Blow-Down Super Charging
(BDSC) system", which can trap a large amount of EGR gas and intake air-fuel mixture without an
external supercharger by using the exhaust blow-down pressure wave from another cylinder phased 360
29
degrees earlier in the firing order, using EGR guide (a circular ring around the exhaust valves and a bank
on the piston were attached to control the rebreathed gas flow), which controls the mixing between the
fresh air-fuel mixture and the high temperature EGR gas recharged through the exhaust port. They
found with EGR guide enhanced thermal stratification, CA50 and CA90 were delayed and combustion
duration became longer, therefore the MPRR was reduced and CAI HLL was successfully extended.
However, CA10 wasn't affected much, probably due to the interaction between fuel stratification and
thermal stratification.
Amir Maria [1] varied coolant temperature to create different extends of thermal stratification
inside engine cylinder for his research on fuel selection for CAI combustion HLL expansion. He found that
to take full advantage of the naturally existed thermal stratification, fuels used shouldn't to have a
predominant Negative Temperature Coefficient (NTC) region, so that the sensitivity of ignition delay to
temperature can be maximized. As shown in Figure 1-3 (reprinted from Maria [1]), three different fuels
were tested on a port injection CAI engine with coolant temperature reduced to 50 *C for the thermal
stratification creation and intake temperature varies for combustion phase adjustment. At the given
operation conditions, they show different knock limited NIMEP (Knock Intensity less than 3 MW/M 2 )
with the same combustion phase, e.g. at
CAMPRR
Of 357 'CA ATDC intake, fuel 2 has the highest load limit,
and fuel 1 and 3 follows in order.
7U
LU
Z
&
0
(U
5
-0-Fuel #2
-Fuel #3
4
.
350
352
354
356
:358
360
362
Location of MPRR [CAD after TC Intake]
Figure 1-3
Knock-limited NIMEP versus CAMPRR; fuel comparison; port injection, MAP
rpm; reprinted from Amir [1]
30
=
1.9 bar, 1900
The insight behind it is the different NTC region behavior of the fuels as shown in Figure 1-4
(reprinted from Maria [1]). It is apparent that Fuel #3 exhibits a strong NTC region where little
temperature sensitivity exists, and the benefit of temperature stratification is diminished. Notice during
the compression process when the in-cylinder charge concentration increases, the NTC region moves to
a lower ignition delay time.
Temperature [K]
Temperature [K]
1000 950
100'
W
E
10
800
850
900
1250 mo/M
(=0.25
3
750
700
1000
650
__ _ _
1181
100
E
10
,
850
900
950
700
750
800
650
100
100
10
10
E
1=
-0
C
ft!
*8O
.2
2
1250 mo/M3,
0=0.25
1
1
1.0
1.00
1.05
1.10
1.15
1.20
1.25
1.30 1.35
1.40 1.45
1.50
1.00
1,55
.011
1.05
1.10
1.15
1000/Temperature [1/K]
1
(a) Fuel #1
.012
.012
0
.
1.2p
.015
.014
1.35 1.4
1.K
000/Temperature [I1/K]
(b) Fuel #2
Temperature [K]
1000 950
900
850
750
800
700
650
100
10
10
E
a
E
1=
a2
0
1250 mol/r3
0=0.25
100 1.0
.1
1.05
1.0.101.2115 130
.15
1.20
/m 1.25 1.30
tu
1.35
115
10
01
14
1,r
15
1
1GOO/Temperature [1/Kj
(c) Fuel #3
Figure 1-4 Ignition delay curve corresponding to (a) Fuel #1, (b) Fuel #2, and (c) Fuel #3; MAP = 1.9
bar, CAMPRR 357 CAD; reprinted from Maria [1]
From above we find that:
1.
Thermal stratification naturally exists but is difficult to create or control, as very unusual
methods is needed to achieve the variation of in-cylinder temperature distribution.
31
2.
Thermal stratification exhibits the potential for CAI heat release relaxation, however
sophisticated control over lots of other operation parameters is required to well utilize it.
1.2.4
Fuel Stratification
The inhomogeneity of ignition characteristics among the mixture parcels in CAI engines can be
achieved by fuel stratification as well: fuel parcels with higher equivalence ratio would likely be ignited
first and lower ones later. This would potentially lead to sequential ignition and then extended
combustion duration and reduced pressure rise rate, eventually improved high load limit. The control of
fuel stratification can be realized by varying the timing of the fuel directly injected into engine cylinder in
the late compression stroke, or by varying the amount of the fuel directly injected into engine cylinder
while the total fuel amount kept constant with premixed charge as compensation.
Sj6berg [40] tested the method of partial fuel stratification for CAI combustion heat release
relaxation in a modified Cummins B-series diesel engine with PRF fuels. Based on the results of a multizone CHEMKIN model and experimental knowledge on how the injection timing influences charge
mixture inhomogeneity, the partial fuel stratification was designed to provide a background equivalence
ratio by injecting in the intake stroke, and a "stratification" adjustment by injecting in the compression
stroke. PRF fuels were chosen as it was demonstrated by a "alternate-fire" test that the auto-ignition of
two-stage fuels have the highest sensitivity to equivalence ratio distribution, so that the inhomogeneity
can be best utilized with such fuels. It was found that stratification effectively increased the IMEP output
from 5.4 bar to 6.0 bar. It was also pointed out that special care is needed to make sure no overly lean
or rich regions that deteriorate the emission benefits from CAI combustion. Sj6berg insured that with
the first injection at 40 CAD and the second at 285 CAD after TC, NOx formation ($>0.5) and CO
formation (4<0.18) were effectively avoided. At the optimal stratification and maximum load,
approximately 17% of the fuel was injected in the second injection.
Dec [41] improved the maximum IMEP of a boosted gasoline CAI engine from 11.7 bar to 13 bar at
intake pressure 2 bar by using partial fuel stratification while CA50 was held constant by varying the
amount of CSP dilution (i.e. some of the excess air is replaced with CSP). It was found that stratification
did not have a large impact under atmospheric conditions. This is saying that the effectiveness of fuel
stratification could be impacted by the ignition and combustion conditions and fuel properties.
As
shown in Figure 1-5, the advancement of the ignition timing (CA10) for homogeneous mixtures with
different equivalence ratio shows very different trend for under atmospheric intake pressure and
32
elevated intake pressures. Therefore with uneven distribution of equivalence ratio, the sequential
ignition would only be better facilitated for those highly sensitive cases. This test was also conducted
with the "alternate-fire" technique to make sure the cylinder wall temperature does not disturb the
results but the results are all chemical kinetic govern.
371
* 376
370
I-
I
375
-
369
-
-
374
C.)n 368
-
-
373
367
-
-
372
366
-
- 371
365
-
364
-
0
3Q3
-- PRF7,Pin=1 bar, base-ph 0.42
-Gasolin, Pin =1 bar, basephl = 042
-4- Gasoline, Pin = 1.6 bar, base-phi = 0.367
-4--Gasoline, Pin = 2 bar, base-phi = 0.345
-4-Gasoline, Pin = 2 bar, base-phi =0.42
,
,
,
.
.I
-
370
- - -
369
.I
U)
0
0-
0
368
0.3 0.32 0.34 0.36 0.38 0.4 0.42 0.44 0.46 0.48 0.5
Charge Mass Equivalence Ratio [[],]
Figure 1-5 Equivalence ratio sensitivity of gasoline and PRF73 under different intake pressures;
reprinted from Dec [41]
Maria [1] investigated the fuel stratification effects on CAI engine HLL expansion by delivering fuel
via single pulse direct injection (DI). He found with single pulse DI in the early compression stroke
(before 110 *CA BTDC compression), sequential ignition is likely to be enhanced and knock limited
NIMEP is increased comparing with port injection. With a bit later Dl timing (110 'CA to 30 'CA BTDC
compression), the effect is diminished because: 1) even heavier fuel stratification causes the
concentrating of fuel in one area of the chamber, and the full chamber temperature distribution is not
used; 2) high fuel concentration would counteract the locally low temperature effect and made the
ignition characteristics more homogeneous among mixture parcels; 3) high fuel concentration will lead
to locally high volumetric heat release rate. With very late injection (after 30 *CA BTDC compression),
the combustion becomes diesel like with low pressure rise rate but very high NOx emissions due to
insufficient time for evaporation and mixing, which should be avoided. He also found the intake
conditions have significant impact on the effectiveness of this strategy.
33
Therefore using fuel stratification to create mixture inhomogeneity has promising potential for CAI
engine HLL expansion, as it's more readily controllable and practical for product engines than thermal
stratification. However its effectiveness is also impacted by lots of factors such as intake conditions,
combustion phase, fuel property, interactions with thermal stratification etc. Also special care needs to
be taken to make sure the extend of the fuel stratification is within proper range that emissions due to
too lean or too rich mixture can be well controlled to retain the benefits of CAI combustion. This
research will systematically study the effectiveness of fuel stratification under different conditions for a
comprehensive picture of how it should be properly and practically utilized, based on the linkage
between the engine performance and fundamental knowledge of the auto-ignition phenomenon.
1.2.5
Other Strategies
Many other techniques and strategies have also been explored for the purpose of CAI combustion
control and HLL expansion. Below is a brief review of some of them for the purpose of completeness of
the scope.
1.
INTAKE TEMPERATURE CONTROL
Intake temperature control is one of the most extensively used techniques on CAI engines. Thomas
and Bengt Johansson [42] from Lund University examined the effect of intake temperature on a
turbocharged CAI engine, with the intake temperature adjusted by an electrical throttle in the bypass
routing to the water-cooled intercooler circuit of the intake system. They found that combustion
stability, combustion noise and engine efficiency have optimums at different intake temperatures and
combustion timings. The usage of EGR and proper valve timing are needed for improvement of the
alignment of these differences. Wildman [34] studied the intake temperature effect on a CAI engine
with both NVO operation and external EGR capability. He found intake temperature can be used to
effectively vary the charge density and temperature, which would then impact the MPRR of the
combustion, and future work should focus on extending the misfire limit so that the combustion can be
retarded for more power output but acceptable combustion noise. Sj6berg [37] lengthened the burn
duration and reduced the Knock Intensity on a CAI engine by retarding the combustion phasing with
reduced intake temperature.
2.
VALVE TIMING
34
Valve timing adjustment provides the potential for CAI combustion control by varied charge
temperature and composition through the residual gas amount, as well as by the effective engine
compression ratio. Widd and Johansson [43] investigated the SI-HCCI dual mode operation on a Volvo
D5 light duty engine equipped with a fully flexible pneumatic valve train system supplied by Cargine
Engineering. An experimental study of SI to HCCI mode switch was performed with early intake valve
closing (EIVC) and NVO. A model-based controller was implemented to stabilize the HCCI operation with
the desired IMEP and CA50. Scaringe [35] researched the possibility of extending the HLL of CAI
combustion on a single cylinder engine with fully variable valve timing capability. Adjusted NVO timing
and duration effectively varied the amount of hot residual trapped inside the engine cylinder. The high
load limit was found always occurs at the misfire limit.
Yun [15] studied the effect of NVO operation with a Fully Flexible Valve Actuation (FFVA) system on
his engine. With combustion phase controlled by a spark plug, stoichiometric conditions at various fuel
loadings were maintained with varied EGR rate. He found by retarding the combustion phase, NVO
operation could be used to extend the load limit. He also found by trapping residual gas with a Positive
Valve Overlap (PVO) strategy, the pressure rise rate could be reduced as a higher portion of the mixture
was ignited with the spark assisted flame.
3.
FUEL SELECTION
CAI combustion can be operated with a variety of fuels, and people have tried to take advantage of
this feature for the HLL expansion as well. Maria [1] fundamentally studied the fuel selection for CAI
engines with the goal of knock mitigation. He found fuel chemical kinetics significantly impact the CAI
engine performance. A procedure was developed to facilitate the design of a CAI combustion system
based on his findings. Hanson et al. [44] investigated the potential of controlling premixed charge
compression ignition (PCCI) combustion strategies by varying fuel reactivity. Excessive pressure rise rate
was successfully prevented on a medium engine load of 9 bar NIMEP by delivering gasoline via port fuel
injection (PFI) portal and diesel via direct injection portal for combustion phasing control. With proper
fuel blend and injection timing, control and versatility of dual-fuel PCCI combustion was demonstrated.
Dec also [41] found that fuel selection could significantly impact the effectiveness of stratification in a
CAI engine.
1.3
Research Motivation
35
A variety of technologies have the potential for CAI combustion control and HLL expansion as
reviewed above, yet three important problems are also seen holding it back from further progress:
1.
The effect and the effectiveness of these methods are not always found consistent.
2.
The mutual impact of different strategies is not clear.
3.
Lots of the methods could be impractical and costly on production engines.
To advance the CAl combustion technology toward industrial practice, more systematic research on
realistic techniques for CAI HLL expansion should be conducted. Fuel stratification and EGR out of all the
above methods showed extraordinary value for production engines besides their potential on effective
CAI combustion control. With the trend of intake boost enhanced engine downsizing booming in the
industry, better understanding on the effect of fuel stratification and EGR dilution on boosted CAI
engines becomes urgent and important. To do that, both on-engine experimental research and
fundamental knowledge regarding the auto-ignition process of the fuel-air-dilution mixture are needed.
1.4
Research Approach
To systematically study the effect of fuel stratification and EGR on CAI engine knock mitigation
under intake boost condition, engine based experimental work needs to be conducted while the
understanding of fuel ignition characteristics and its impact should be used for the engine performance
interpretation. The research project was carried out in three steps as outlined below.
1.4.1
Step I - Fuel Characteristics Investigation
Tests on rapid compression machine (RCM) are conducted to obtain the ignition and heat release
characteristics of the specific fuel used for the research by measuring the ignition delay and MPRR of the
fuel-air-diluent mixture under different initial temperatures. The effect of air dilution and inert gas
dilution are compared for better understanding of the effect and relative importance of each method.
Different diluents are used to isolate the chemical effect of the mixture components from each other.
Fixed charge density tests are conducted to demonstrate the sensitivity of the ignition delay to
equivalence ratio distribution, and fixed fuel density tests are carried out to show the dilution effects of
the inert gas and air under comparable conditions on boosted CAI engines.
1.4.2
Step II - Engine Tests with Fuel Stratification and EGR
36
Engine experiments are conducted to map out the effect of fuel stratification and EGR on CAI knock.
The working range of the engine under boosted conditions is obtained first with premixed charge.
Proper test points are then identified and picked for fuel stratification and EGR study. The DI injection
timing and DI injection ratio are varied for fuel stratification adjustment. EGR rate is varied for all the
injection timings. Different compression ratios are tested to demonstrate the difference of the
effectiveness of fuel stratification. The engine operation conditions (speed, intake temperature) are also
varied for the purpose of comprehensive research.
1.4.3
Step III - Engine Performance Interpretation
The engine performance are analyzed and interpreted with the understanding of the fuel autoignition process and the fuel characteristics on ignition and heat release obtained before. The
effectiveness of the strategies is illustrated and guidelines for effective usage of the fuel stratification
and EGR on CAI engine HLL expansion are proposed.
1.5
Research Objectives
The specific objectives of this research are:
1.
Obtaining the fuel auto-ignition characteristics under different concentration, dilution, and
temperature corresponding to in-cylinder conditions of typical high load CAI combustion.
2.
Demonstrate the fuel stratification and EGR effect on knock mitigation in boosted CAl
engines.
3.
Connecting the chemical kinetics fundamentals with the engine performance for a
comprehensive and through understanding of the effectiveness of fuel stratification and EGR
under different operation conditions.
1.6
Report Outline
Chapter 2 describes the experiment systems and apparatus used for the research. The specifications,
working procedure, and typical data of a RCM built in lab and a production engine based testing system
are introduced.
37
Chapter 3 reviews the physics and the metric of CAI engine knock. Different theories of explaining
the phenomenon are introduced and reasoned. The knock metric of this research was determined based
on careful analysis and comparisons.
Chapter 4 shows the research regarding the fuel ignition characteristics. Experiment results for
mixtures with fixed fuel density and charge density are presented respectively. Tests with different
diluents and duel density are compared.
Chapter 5 presents the CAI engine responses to different strategies of knock mitigation. Knock
performance and combustion performance of the engine under different operation conditions are
reported and compared.
Chapter 6 analyzes the different engine responses based on the understanding of the fuel
characteristics and the auto-ignition
process. The linkage between
engine
performance
and
fundamentals are established. Applications of the findings to future CAI combustion system are inferred.
Chapter 7 concludes the work with the key findings and ideas of future work.
38
Chapter 2. Experiment Apparatus
The experimental part of this research is conducted in two parts: fuel characteristic study on RCM
and engine combustion behavior on a modified production engine. The purpose of the RCM
experiments is to understand the ignition and heat release characteristics of the specific fuel used in this
research. The purpose of engine tests is to systematically examine how fuel stratification and EGR
dilution impact the CAI combustion knock propensity.
This chapter presents the working principle and specifications of the experimental apparatus used,
and the organization of the experimental tests.
2.1
Rapid Compression Machine
A Rapid Compression Machine is an instrument designed to study fuel characteristics by
simulating the compression and combustion process of a single stroke piston-cylinder apparatus.
Comparing with engine tests, the operation of RCM minimizes the variation of impacting factors such as
pressure, temperature, and composition, so that it allows the study of ignition characteristics of fuel-air
and dilution gas mixtures under well-controlled conditions. One combustion event occurs per test run
on the RCM.
For this research, RCM is the most suitable device comparing with other combustion research
devices. Static reactors are designed particularly for low temperature ignitions (typically lower than 750
K). Flow reactors are designed for chemical kinetic interpretations. Well-stirred reactors are for steady
state constant pressure combustion researches. Shock tube, although it is designed for achieving autoignition characteristics of gas mixtures at higher temperatures and pressures, typically has much shorter
test duration (~less than 5ms) than RCM due to interference from boundary layer effects and/or
reflected waves. Comparing with these devices, RCM gives accessibility to study combustion with
ignition temperature from 600K to 1100K, which is perfect for this research according to the typical
ignition temperature range of CAl combustion on engine, the given test engine and the fuel used. The
typical test duration on RCM is around 100 ms, which gives it great capability to cover ignition tests in
lower temperature regime and profile a full picture of the fuel ignition characteristics [45].
39
The RCM used in this research is a vertical machine built by Amir Maria [1] during his study at MIT as
shown in Figure 2-1. It consists of five major parts: core part, mixture preparation unit (MPU), sampling
apparatus, hydraulic system, and auxiliary components (gas tanks, heating systems, and DAQ systems
etc.). The RCM is capable of giving quasi-isentropic compression to the bulk part of the fuel-air mixture
in the center of the RCM combustion chamber. This is achieved by shooting the compression piston up
with high pneumatic pressure and let it rapidly reach the top dead point from the bottom point in less
than 15 milliseconds. After that, the position of the piston interfacing with the combustion chamber is
held still by very high hydraulic pressure until the ignition and combustion process completes. A
piezoelectric pressure transducer installed in the combustion chamber reflects the pressure history
during the test. The pressure trace is then used to determine the ignition delay time and profiling the
heat release process.
Figure 2-1
2.1.1
MIT vertical RCM
RCM Specifications
Table 2-1 shows the key specifications and dimensions of the RCM, as well as the limitations of its
key components. These values are important references for the design of experiment tests.
40
Table 2-1
2.1.2
RCM design specifications and component limitations [46]
Combustion Piston Diameter [m]
0.0508
Stroke [m]
0.2032
Compression Ratio
8,10,12,14,16,18,20
Maximum Combustion Pressure - no sampling [bar]
200
Maximum Combustion Pressure - sampling [bar]
80
Pneumatic Piston Diameter [m]
0.127
Pneumatic Pressure [psi]
250
Hydraulic Pressure [psi]
2000
Maximum Initial RCM Pressure [torr]
2000
Maximum Initial RCM Temperature [*C]
100
Maximum Mixture Preparation Unit Pressure [bar]
3
Maximum Mixture Preparation Cylinder Temperature [*C]
80
Maximum Mixture Preparation Network Temperature [*C]
60
MIT Vertical RCM Working Procedure
1. Mixture Preparation
A RCM test starts with mixture preparation that is carried out in the MPU (the cylinder with stainless
steel top, black insulation jacket and extruded tube-valve sets in the left side in Figure 2-1). The
operation procedure can be illustrated with Figure 2-2 [46].
Gas Bottle
Gas Bottle
Viso
V19
VISA
V19A
Gas Bottle
V208
V2
O
100 tor
Mixture Chamber
3 bar
2000 torr
h
V12
Vi'
Vac u
1A
To RcM
S Inlet
Liquid Fuel Inlet
Vacuum
Vent
Figure 2-2 Mixture preparation unit (MPU) schematic [46]
41
The way MPU works is based on the partial pressure law of ideal gases. Calculation is conducted to
obtain the partial pressures of each species in the MPU cylinder for specified mixture composition, and
different gases are admitted into the cylinder in order while the accumulated pressure is monitored
from the pressure meters (P4 and P5).
The MPU chamber is heated up to certain temperature (typically 60 "C) by the electric heating sticks
inserted into the bottom and the top to prevent condensation of fuel and water vapor according to the
mixture pressure. With all the other valves closed, the mixture chamber is evacuated by opening V17,
which connects to a vacuum pump. V15 and V16 are opened to observe the pressure in the chamber
while pressure meter P4 associated with V15 is specifically for the fuel pressure observation with an
upper limit of 100 torr. The fuel is injected with a syringe from the liquid fuel inlet by an amount
calculated beforehand with the maximum pressure of the mixture chamber is 3 bar kept in mind. After
the pressure in cylinder stabilizes with only fuel vapor, V15 is closed. The valves in position 18, 19, and
20 are connected to gas bottles to provided needed oxidizer and dilution gas. In this research, high
purity oxygen and nitrogen is used for the nitrogen dilution tests, while carbon dioxide and water vapor
are added for the EGR simulation tests. Notice the water vapor is delivered by syringe through the liquid
fuel inlet as well. In the portals connecting to gas bottles, "A" valves are flow rate adjusting valves and
"B" valves are open/close valves for the purpose of override. With the partial pressure of each species
and the total accumulated pressure calculated beforehand, the valves connecting to gas bottles are
opened and closed one after another in certain order to achieve the target equivalence ratio and
dilution rate by watching the pressure from meter P5. After all the elements are delivered into the MPU
cylinder, the fan on top will be turned on for 10 minutes to help the mixing process occur faster for a
homogeneous mixture. With the composition and temperature of the mixture in MPU stabilizes, it can
be admitted into the RCM combustion chamber through V13 and 11. The meter P6 is used for watching
the pressure of the mixture in the RCM combustion chamber to make sure the amount of gas admitted
works together with the RCM geometry and temperature to give the requested charge and fuel
concentration. V14 is used to evacuate the RCM combustion chamber before the test mixture is
admitted into it as well as after the test is finished.
2. RCM Operation
With mixture prepared in the MPU, test runs can be completed in the RCM core part by utilizing the
hydraulic network for pressure control. The core part can be seen as the stainless steel tower in the
42
middle of Figure 2-1. Figure 2-3 shows the configuration of it. The operation procedure can be illustrated
in Figure 2-4 with the hydraulic network affiliated.
From Figure 2-3 we can see that the core part of MIT vertical RCM mainly consists of three
cylindrical structures and one piston shaft set----the moving assembly of the RCM. From the bottom up,
the pneumatic pressure access is connected with high-pressure nitrogen (250 psi) to provide driving
force to the piston shaft set for the fast compression motion by pressing the downside of the pneumatic
piston. The pneumatic chamber restraints the movement of the pneumatic piston in the vertical
direction and is open to the air. Connecting via the pneumatic shaft, the hydraulic piston moves with the
piston shaft set and seals the high-pressure hydraulic oil in the hydraulic chamber by its top-facing side.
Above it, the combustion piston seals the combustion chamber together with a top cap and contains the
fuel-air mixture admitted from the MPU. The pin and groove mechanism [46] makes sure of crash-free
contacting of the hydraulic piston with the top and bottom of the hydraulic chamber at their extreme
dead points. (When the annular pin in the hydraulic piston enters the groove in the hydraulic stop ring,
large viscous forces generated in the tight clearance between the pin and groove slow down the moving
assembly to approximately 5m/s as the hydraulic oil is highly incompressible.)
Fast Sampling
Access Points
Combustion Chamber
Combustion Piston
Pin and Groove
Oil Pressure Access Line
Mechanism
Hydraulic Chamber
Oil Fill Access Line
Hydraulic Piston
Pneumatic Chamber
Pneumatic Shaft
Pneumatic Piston
Pneumatic
Pressure Access
Figure 2-3 RCM core part [46]
To start a RCM test run, the moving assembly is drawn down to the lowest point by opening V6 to
vacuum, as shown in Figure 2-4. With the mixture admitted into the combustion chamber, a 10 minutes
time period is needed before firing for a homogeneous temperature of the mixture inside. Then the
43
hydraulic piston is sealed against the stroke adjustment ring with the bottom vented by open V3, V6 to
vacuum, and V8 to vent. After the moving assembly in position, the hydraulic oil in the chamber is
pressurized to 2000 psi by open V1 to the high-pressure nitrogen. The pressure in the hydraulic chamber
is read from P1. When the pressure stabilizes, V3, V6 and V8 are turned off to isolate the hydraulic
network and the RCM. The pneumatic pressure access is then through to provide the 250 psi pressure
for the pneumatic piston. When the DAQ system is ready, V2 is opened to "fire" the RCM as this releases
the high pressure in the hydraulic chamber in a flash and lets the moving assembly shoot up rapidly.
Notice although the area ratio of the pneumatic piston to hydraulic piston is as high as 6:1, the high
pressure in the hydraulic chamber could hold the moving assembly firmly before V2 is turned open. The
combustion piston will be hold on at the top dead point until the heat release process is completed.
Vent
Vent
Vent
High
Pressure
N2
VI
lIPRP
Vz
'a'
Oil/N 2
Aux.
Vs
Mi .............
I
DraidF
Vacuum
Vent
Vent
Ns Vent
Pneumatic
Chamber
Pump
Closed
Oil Reservoir
Drain
Figure 2-4 RCM hydraulic network schematic [46]
2.1.3
Test Condition Set-up and Data Acquisition
The set up the ignition conditions for specified mixture composition, the RCM initial temperature,
compression ratio, and the initial pressure/molar concentration are varied. These values are achieved by
44
calculating the after compression temperature with the RCM geometry and the target charge
concentration. The calculation is based on the assumption that the core part gas inside RCM combustion
chamber is compressed isentropically (adiabatic core hypothesis) [1]:
dT =
Tcomression
fTi
~nta
Cp(T ) initial=Ru In
Tinioi
t i1
[2.1]
Pcompression
The in-cylinder pressure is recorded during the test by a Kistler 6125A piezoelectric pressure
transducer plugged into the combustion chamber. A Kistler 5010 charge amplifier is coupled with the
transducer for converting the electric charge information into a voltage signal. The voltage signal is then
filtered and post processed to obtain the pressure trace. The pressure trace is finally used for the
ignition delay and heat release rate determination, which is discussed in further details in Chapter 4. NI
USB-6211 DAQ device is used for data acquisition. With the ability of maximum multiple channel
sampling rate 250 KS/s, a 10K HZ sampling rate is set up for the cylinder pressure signal here with a lab
made LabVIEW Program.
An example of the test data is shown in Figure 2-5.
140-
120
100
CO 80
Phi= 0.6
Dilution=0.4
40
Tinmtal=32 *C
201
00
0
10
15
25
20
30
30
40
4
Time [ms]
Figure 2-5 Pressure traces from the RCM tests
Easy to see that the compression ratio is varied to get different ignition temperature for the same
mixture, while the initial temperature is fixed at 32 *C in this particular test set. The initial temperature
45
is controlled by the thermal jacket around the RCM combustion chamber wall and the electric heat
sticks inserted into the combustion chamber base, as shown in Figure 2-1. Notice each test condition
corresponds to two pressure traces. This is to make sure the test results are repeatable. A maximum
error of 10% difference of the ignition delay and maximum pressure rise rate between the two test-runs
is set as the tolerance.
2.2
Engine Experiment System
Engine based experiments are designed in this research to test the practical effect of fuel
stratification and EGR on high load limit expansion for boosted CAI combustion. Engine operation
conditions and the extent of fuel stratification and EGR dilution are also varied for comprehensive
understanding of the effects and factors impacting their effectiveness. The engine experiment system is
shown in Figure 2-6.
Figure 2-6 Engine experiment system
The test system is built upon a modified 1.9L Renault Laguna diesel engine, taking advantage of its
original high compress ratio for feasible CAI start up, and its sturdy structure for operation under knock
conditions. The gas charging system was enabled with intake boost, exhaust pressure control, and EGR
46
variation. Electric heating systems are set up for the intake gas and recycled exhaust gas to well control
the initial charge temperature. Two independent fuel delivery systems, PFI and DI, are configured in
parallel to operate simultaneously or separately for the purpose of fuel stratification creation and
adjustment. The amount of fuel injected is controlled in real time by voltage pulse command sent from
the master computer. A Leeson Wattsaver 10 hp electric motor drives the engine, and an AE 150 HS
dynamometer loads the engine. The motor power is controlled with a Durapulse 15 hp motor speed
controller.
With controls available for operation parameters, measurements and feedbacks are obtained with
sensors and instruments. A piezoelectric transducer is set up for in-cylinder pressure measurement. The
concentration of C0 2, H20, and HC in exhaust and intake gas can be obtained from an emission analyzer.
A lambda-NO sensor is used for excessive air coefficient and NOx emissions measurement.
2.2.1
Engine Specifications
The original engine
(Renault F9Q B800 common rail diesel engine) is shown in Figure 2-7. To
eliminate cylinder variation issues and operate in single cylinder mode, only the first cylinder is fired. All
the other three cylinders are motored when the first cylinder fires. Related modifications to the intake,
exhaust and fuel system are described below.
Figure 2-7 Renault F9Q B800 common rail diesel engine
47
The engine specifications are obtained from the Mitsubishi F9Q series manual [47] and given in
Table 2-2.
Table 2-2
Description
Specification
Number and arrangement of cylinders
4 in-line
Total displacement
1870 cm 3
Cylinder bore x stroke
80 mm x 93 mm
Compression ratio
19
Valve mechanism
Single overhead camshaft
Valve timing
Intake opening
BTDC 30
Intake closing
ABDC 210
Exhaust opening
BBDC 460
Exhaust closing
BTDC 60
Fuel injection system
2.2.2
Engine specifications [47]
Direct injection common rail (rail pressure: 450 - 1500 bar)
Intake and Exhaust System
The intake and exhaust runners were recreated separately for the firing cylinder, and the original
manifolds were modified to connect to only the non-firing cylinders, as shown in Figure 2-8.
Figure 2-8
Modified intake and exhaust system
48
To enable the engine operate in elevated intake pressures without being ifluenced by the operation
conditions, the original turbocharger affiliated to the engine was removed, and a Mini
Cooper
supercharger was installed, driven by a Marathon Electric 10 hp motor. This allows the engine
operates
with intake pressure from 1 to 2 bar (absolute pressure). An intake piping system was also
adopted from
Amir Maria [1] to facilitate the intake pressure control as shown in the left hand side of
Figure 2-9. We
can see that after the fresh air is drawn into the first damping tank, it is compressed
through the
compressor and flows into the second damping tank after the intercooler.
By opening the valve
between the two damping tanks, the pressure in the first damping tank increases, which will reduce
the
flow of the fresh air into the damping tank, therefore reduce the air flow into the supercharger,
and
consequently the intake pressure.
EGR Heat Exchanger
C02
Air in
EGR
msrmn
Damping
Tank
Damping
Tan k
system
ia
Heater
Exhaust
ENGINE
Dampin g
Controlled Tin
nrcoler
MiniCooper
compressor
Bypass
Hae
Hontr
throttlecnto
controllotro
st
throttle
control
Controller
Figure 2-9 Engine intake system schematic
A Kurz 505 thermal anemometer is used for air mass flow measurement by determining the level
of
current required for a velocity element heated at the same temperature [48]. The intercooler used
could
bring the temperature down to approximately 30 'C, depending on the desired air pressure. The valve is
controlled by an intake stepper motor, which is driven by a Danaher Motion P70360
micro stepping
motor drive. The power supplied to the motor driving the supercharger (controlled by a Durapulse 10 hp
motor controller) could also be varied to adjust the outlet pressure of the compressor, however the first
method is desirable because it gives the airflow an outlet in case the engine stalls [1]. Out of the second
49
damping tank, the air flows through a Sylvania 6 kW heater, and the heated and pressurized air is then
directed to the intake of the firing cylinder. The intake temperature can be adjusted in the range from
40 "C to 140 C. An Avatar A3Z power controller coupled with an Omega CN77000 series controller
varies the power supplied to the heater for the purpose of intake temperature control. The intake
temperature is measured on the intake runner connected to the firing cylinder and used as feedback for
the controller.
Out of the engine, the exhaust is passed to a third damping tank, as well as an EGR loop. The third
damping tank is connected to the laboratory trench. The exhaust valve between the third damping tank
and lab trench is used to adjust the exhaust pressure and thereby facilitate EGR for the engine under
boosted intake pressures. As the exhaust valve closes, the backpressure increases, leading to a higher
exhaust pressure. Throughout the research, the exhaust gas pressure has always been adjusted at 0.1
bar higher than the intake. An exhaust stepper motor is used for valve actuation, and it is driven by a
Danaher Motion P70360 micro stepping motor drive.
The EGR rate is adjusted via a Proportion-Air electro-pneumatic proportional flow valve, which is
controlled by 0~10V voltage signal from a lab made potentiometer and is driven by compressed nitrogen
at 80 psi. The valve employs a parabolic valve plug so that the area of valve opening is proportional to
valve position. This greatly helps achieve precise and reliable controlling of EGR rate. The valve is
designed to work in environment with temperature from 0 0C to 180 *C and pressure up to 17.25 bar,
which could well suit the requirement of after-cooling exhaust gas.
A Volkswagen TDI EGR cooler (labeled as "EGR Heat Exchanger" in the figure) is installed between
the EGR valve and engine exhaust runner to cool down the high temperature exhaust gas before it mixes
with the fresh charge. City water was supplied as coolant for the heat exchanger, and the flow rate is
adjusted by a Parker on/off solenoid valve. An OMEGA flexible heating tape SRT101-100 is winded
around the gas pipeline right out from the EGR heat exchanger. Two OMEGA CN7800 controllers are
used to control the on/off action of the solenoid valve and the heating tape separately. The gas
temperature is measured right out of EGR valve and before the intake heater joint respectively as
feedback signals of the closed loop temperature control. A Horiba MEXA-554J automotive emission gas
analyzer was used to measure CO 2 concentration in the intake and exhaust gases for the purpose of EGR
rate calculation. An even simpler method used for EGR rate bench marking in this research is by
comparing the fresh air mass flow rate (converted into standard conditions: 25 "C and latm) under EGR
50
operations and no EGR operation, which turned out to be more effective and sufficiently accurate. This
is also saying that potential leaking along the intake and exhaust pipelines is negligible.
2.2.3
Fuel Delivery System
The fuel delivery system consists of two subsystems: direct injection system and port fuel injection
system. The two subsystems could work either independently or in collaboration.
1. DI System
The DI system of the firing cylinder was modified from the original diesel engine fuel injection
system. With test fuel contained in a special tank, an engine driven fuel pump draws the fuel from the
tank through a fuel filter. The fuel then either bypasses away from the fuel rail (a large portion) or gets
pressurized to maximum pump pressure and fed to the fuel rail, depending on the position of a control
solenoid valve. The solenoid valve is driven by a PWM controller, which varies its duty cycle in the
manner of closed loop control based on the desired fuel pressure and the actual fuel pressure. Fuel in
the common rail can then be directed to the direct injector of the firing cylinder, which is adopted from
the original product engine. A custom injector driver, which produces the opening and holding current
based on the desired injection timing and duration, is used to control the fuel injector. The control signal
is sent out from the master computer in real time (effective from next cycle on) with a minimum
available pulse width of 200 ms. Since expansion from high pressures of gasoline can lead to flash boiling
and alter the fuels composition in the case that fuel components do not condense before re-entering
the fuel tank, a heat exchanger is set up to bring the fuel temperature down to the temperature of the
city water supply (10 *C to 20 'C) [1].
Notice the backpressure to the injector nozzle (in-cylinder pressure) varies with the injection timing,
which might cause variation to the actual amount of fuel delivered into the cylinder. From injector
calibration data we know that with an injection pressure at 500 bar, there is no countable changes
noticed in the majority cases (Start of Injection (SOI) earlier than 160 'CA ABDC intake), and the change
of the total amount of fuel is less than 5% for cases with SOI later than 160 'CA, which wouldn't
substantially affect the engine knock performance. For example, Figure 2-10 shows the fuel delivery
distribution for testes with different DI timing and EGR rate under the condition of CR=15, RPM=1200,
intake temperature=80 0C, total amount of fuel around 25mg, and PFI/DI=40/60. Easy to see that for the
51
DI timing range interested the most (30*CA to 150*CA), the total amount of fuel was well kept constant,
and even too late DI SOI wouldn't cause substantial drop.
Very slight difference for very late DI SOI cases
25
t
20
15
III[
m
JilL
mi
" PFI Mass (mg)
" DI mass (mg)
10
5
-
-
11111
45 60 7-5 90 1051201301501601
EGR =0
O
30 45 60 75 90 105120135150160170
EGR =10%
0 45 60 1201351s
16s30
EGR =15%
liii
i
DI SOI
~
s1401Ol
6
j
(deg CA
ABDC intake)
EGR =30%
(total amount difference may be
due to different weather in another day)
Figure 2-10 Fuel delivery distribution for varied DI SOI
2. PFI System
The port fuel injection subsystem is set up by placing a Bosch EV6 PFI injector on intake runner of
the firing cylinder toward the intake valve. The injection pressure is set at 40 psi and is provided by SI
engine standard fuel pump and pressure regulator. A Global Specialties 4001 Ultravariable Pulse
Generator controls the amount of fuel injected per cycle, with the pulse width manually adjusted and
the timing triggered by BDC signal from a Litton Rotary Shaft Encoder. The amount of fuel injected can
be consistently varied in a large range of control pulse width under different intake pressures as shown
in the PFI injector calibration curves in Figure 2-11.
52
0
18
2000
4000
6000
8000
10000
12000
118
16-
- 16
14-
14
I....
a,
12
12
-
10-
10
E 8-
8
C-)
0)
c..J
0~
a,
6-
6
M AP=2.76 bar (Intake P=1 bar)
4-
I
U
a,
*0
*
M
.
0
20-
0
2000
4000
AP=2.56
AP=2.36
AP=2.16
AP=1.96
AP=1.76
6000
bar (Intake
bar (Intake
bar (Intake
bar (intake
bar (Intake
8000
P=1.2
P=1.4
P=1.6
P=1.8
P=1.9
10000
bar)
bar)
bar)
bar)
bar)
- 4
- 2
- 0
12000
Pulse width (ms)
Figure 2-11 PFI injector calibration chart
The returned fuel is first cooled down through copper coils merged in ice water before enters the
fuel tank. The system is shown in Figure 2-12.
Figure 2-12 Pressure supply and fuel return of the PFI injection system
53
With the above system, the amount of fuel injected by DI and PFI is well controlled with the
injection pulse width and is verified with the calibration charts. The fuel ratio could be changed feasibly
according to test requirement as well. A Horiba MEXA-720 NOx analyzer is also used to monitor the
excessive air coefficient and thereby bench mark the total fuel mass injected.
2.2.4
Engine Controls and Data Acquisition
Controls over different operational variables of the engine are summarized by Amir Maria [1] and
added on by the author in Table 2-3.
Table 2-3
Control
Variable
Engine Speed
Inlet Pressure
Exhaust
Pressure
Coolant
Temperature
Inlet
Temperature
Sensing
Engine control variable summary
Instrument
Dynamometer Speed
Pick-Up
Honeywell 9301202
Absolute Pressure
Transducer
Omegadyne Absolute
Pressure Transducer
Thermocouple
Method of Control
Power input into electric motor driving engine and dynamometer. The
motor is loaded at a fixed load by the dynamometer. Motor power is
adjusted to attain desired engine speed.
1. Position of valve connecting damping tanks 1 and 2
2. Power input into electric motor driving supercharger
Position of valve connecting exhaust damping tank (damping tank 3) to
trench
Rod heater power and solenoid valve controlling cooling water into
heat exchanger [closed loop control]
Thermocouple
Power input into intake heater [closed loop control]
DI Pressure
Bosch Pressure Sensor
DI Timing
N/A
Duty cycle into PWM driver connected to fuel pump solenoid valve
[closed loop control]
Direct input into custom injector driver
DI Duration
N/A
Direct input into custom injector driver
PFI Pressure
PFI Timing
Pressure regulator
N/A
Standard fuel pump and pressure regulator
Direct input and BDC trigger to Pulse generator
PFI Duration
N/A
Manual adjustment on pulse generator
EGR
Horiba MEXA -554J for
Exhaust gas flow into intake throttled by proportional valve with 0-10V
CO
DC voltage
2
A NI USB-6211 DAQ device gathers all the feedback data with the maximum sampling rate for
multiple channels of 250KS/s.
A lab made LabVIEW program is used for the data monitoring and
recording.
54
The NOx emissions measurement becomes obtainable with the Horiba Mexa-720NOx Analyzer
mentioned before. Crank Angles and BDC signals are counted by a Litton Rotary Shaft Encoder on the
engine front. The most important feedback signal----cylinder pressure is achieved by a Kistler 6055
piezoelectric pressure transducer mounted from the original glow plug position of the firing cylinder,
coupled with a Kisler 5010 charge amplifier.
The pressure data is acquired in two different manners: "slow" and "fast". The slow data is sampled
every 1 degree crank angle (7.2 KHz at 1200rpm) for 100 cycles and is used to analyze the engine
combustion behavior. The fast data is sampled in a frequency of 100 KHz for 3 seconds (30 cycles at
1200 rpm) and is used to capture the engine knock performance. For the fast data, a band pass filter of 4
to 20 KHz is employed to obtain the amplitude of the pressure oscillation, which is defined as the cycleaverage of the maximum absolute value of the pressure amplitude in each cycle of the filtered pressure
trace. The MPRR of the pressure data is obtained from numerically differentiating the low-pass filtered
(below 4 KHz) pressure trace. The cycle-averaged MPRR is then the average of the MPRR of each cycle.
Figure 2-13 shows the fast data of 10 consecutive cycles under knocking condition and the
corresponding pressure oscillation obtained from the bandwidth filter, copied from Maria [49].
18
08
16
06
U0)
a)
142 -_
0
-0.2
0.
0.6
1
4 -
a) -0.6-
2
~LL-0,8
16
2
2,5
3
3.6
4
465
6
0
065
1
1.5
2
2.5
3
3.5
4
Time [ms]
Time [ms]
Figure 2-13 Pressure traces and pressure oscillation under knock condition [49]
55
4.
5
(this page intentionally left blank)
56
Chapter 3. Nature of CAI Engine Knock and Its Mitigation
To understand the fundamental logic of using fuel stratification and EGR for CAI combustion knock
mitigation, this chapter briefly reviews the physics of the CAI engine knock, the metric of this
phenomenon, and the rationales behind the fuel auto-ignition and heat release control. Notice the
purpose of this review is not thoroughly drilling down to the sciences but rather straightening up the
theoretical considerations of the research before jumping into the experiments.
The Physics of CAI Engine Knock
3.1
3.1.1
Thermodynamics of In-cylinder Pressure Oscillation
As the nature of CAI combustion, the heat release process is not limited by the heat or mass transfer
rate
but rather the chemical kinetics. With extremely homogeneous (both temperature and
concentration distribution) mixture, the ignition all over the space can be completed within a time
period in the order of 0.1ms on CAl engines, while the burn duration of conventional engines could take
up to tens of milliseconds due to the flame propagation governed combustion [1]. In reality, there is
always inhomogeneity among the combustible parcels, and parcels with different temperature and
equivalence ratio ignite at different times. Yelvington [27] and Maria [49] pointed out that when the
local thermal expansion due to combustion becomes so fast that it exceeds the limit of acoustic
expansion rate, the local pressure will be higher than the chamber pressure and couldn't be equilibrated
in negligible time. Such a non-uniform pressure distribution results in the formation of pressure wave at
the resonant frequencies of the chamber. Andreae [4] derived the criterion of local heat release rate for
pressure oscillation development by considering the heat release process of a combustible mixture
sphere with radius r and volume V.
For idea gas with constant properties with equilibrated pressure in the chamber, we have below
equation for the heat release process:
PAV = mRAT
From the
1 st
law of thermodynamics, in time At, we have:
57
[3.1]
mCpAT = 4VAt
[3.2]
Using C, = yR/(y - 1) to combine the two equations and we get:
AV
At
y - 14V
y
[3.3]
P
For acoustic expansion with speed of sound c, we have:
A = 4wr 2 c
At
[3.4]
Combining [3.3] and [3.4], the acoustic expansion would not be able to accommodate the thermal
expansion to maintain uniform pressure when:
.
yP
C
y - 1r/3
[3.5]
Easy to see that the ambient pressure in the chamber, mixture composition, temperature in the
chamber (impacting the speed of sound), and charge stratification (impacting the characteristic length of
combustion) all impact the threshold for pressure wave formation. With higher underlying pressure, a
more severe higher heat release process can be accommodated. With smaller combustion characteristic
length, i.e. more stratified charge, more fuel can be delivered into the cylinder before pressure
oscillation occurs. Also with higher EGR rate, the specific heat ratio can be smaller, and the speed of
sound will be lower as well, so that the criteria for uneven pressure distribution will be loosen up more.
It could also be inferred that since CAI combustion normally has larger combustion characteristic length
than SI engines (end gas auto-ignition in a small region causes knock), with the same level of heat
release rate, it's naturally more prone to pressure wave development.
3.1.2
Acoustics of the CAI Engine Knock
With pressure oscillation formed inside engine cylinder, the engine structure will be excited and
vibrates, and eventually result in acoustic radiation perceived as knock noises.
58
By assuming the oscillations are acoustic in nature (amplitude of the pulsations are small relative to
mean pressure within the cylinder), C.S. Draper [50] first gave out the solution to predict the types of
pressure waves in the combustion space of an circular cylinder with flat ends at right angles to the
cylinder axis by solving the wave equation:
0 2 p2P
t=
[3.6]
c2
2
P
where P is the cylinder pressure and c is the speed of sound. For ideal gas, it can be calculated with:
c=
[3.7]
yRT
The equation is solved under the boundary condition that gas particle velocity normal to the wall is
zero:
ap
-=
an
[3.8]
0
When the cylinder clearance height is small compared to the bore, the oscillation frequencies are
expressed as in Equation [3.9].
C
fin=
am~n
B
[3.9]
where fm,n is the resonant frequency for the m,n vibration mode, am,n is the corresponding wave
number determined from Bessel functions, m and n are the numbers for circumferential pressure nodes,
respectively, and B is the cylinder bore. A uniform temperature distribution is assumed so that the
speed of sound is independent of location within the cylinder. The frequencies calculated by Andreae
are exhibited as an example as shown in Figure 3-1 [4].
59
Figure 3-1
(m,n)
1,0
2,0
0,1
3.0
1,1
2,1
Pm,n
1.84
3.05
3.83
4.20
5.33
6.71
f,,,(kHz)
5.01
8.31
10.43 11.43 14.51
18.25
Frequencies of the in-cylinder gas drum modes of oscillation; reprinted from Andreae [4]
Notice the Pm,n in the table is the same as am,n in Equation [3.9]. The calculation was based on a 2.3
L 14 engine under CAI combustion conditions. Specific parameters used includes: B = 87.5 mm,
y = 1.22, R = 302 J/Kg * K, and T = 1500 K.
With frequency spectrum--amplitude data measured by a knock analyzer for different levels of
engine knock for both SI and HCCI combustion, Eng [28] found the lower frequencies (4KHz-8KHz) knock
dominates over higher frequencies (8KHz-2OKHz) on both HCCI and SI engines. Since the pressure
pulsations in HCCI combustion at the
1 st
mode frequency are on the order of 6 to 10 times as large as
those on knocking SI engine without an undue increase in combustion noise, Eng concluded it's the
higher frequencies vibration essentially causes engine knock. He further supported his conclusion with
the observation that a much larger portion of energy is contained in the high-order modes with SI
knocking combustion as compared to HCCI combustion, as shown in Figure 3-2.
20
-
18S
Slope =0.35
16
E
2
14A
12
7
10 0
IN
HCCI
0.12
SSlope=
6
/
IC
4. -
r
7oe
pecrmD
s
0
0
20
40
60
80
100
120
140
160
4.-8 k~k Power Spectrum Density
Figure 3-2 Power spectrum density distributions for first circumferential mode (4-8 kHz) and higher
order modes (8-25kHz) for SI and HCCI operation; reprinted from Eng [28]
60
By measuring the vibration of the engine block, Eng found the low frequencies oscillation is much
attenuated by the engine structure while the high frequencies oscillation can still be largely reflected
from the block vibration. Based on that, Eng recognized the IC engine knock a phenomenon that the
high frequencies vibration transmitted through the engine block while low frequencies pressure
oscillation being attenuated and dissipated by the engine structure [28].
Differently, Andreae [4] collected both the audible data with a microphone fixed above the firing
cylinder, and the pressure data with a piezoelectric transducer probing into engine cylinder on a CAI
engine. He found similarly the lower frequencies vibration takes the majority portion of the vibration in
both audible data and pressure data for different levels of engine knock, although the specific
percentage is a bit lower for low frequencies vibration in the microphone signals than in pressure
signals, as shown in Figure 3-3. Since the lower frequencies vibration dominates in both audible and
pressure signals, Andreae argued that the source of CAl engine knock should be the lower frequencies
pressure wave.
Ui 5000
_4000
3000(A
60000
58-20 KHz spectral
power
04-8 KHz spectral
power
_40000
8-20 KHz spectral
power
M 4-8 K1Hz spectral
power
-
2000-
20000-
S1000-V
00
IL 0
Heavy
Knock
Slight
None
0
Heavy
Knock
Slight
None
Figure 3-3 Spectral powers from pressure transducer and microphone signals of the four subjectively
classified knock cases. The 1st order mode is for the 4-8 KHz range; the higher order mode
is for the 8-20 KHz range; Reprinted from Andreae [4]
By collecting the microphone and pressure data at the same time, Andreae was able to compare the
time dependence of both. As shown in Figure 1-1, the amplitude of the pressure oscillation on the
cylinder pressure trace decays monotonically with time while the acoustic radiation appears to be wave
packets. No direct interdependence is shown at the same time between the amplitude of the acoustic
radiation and that of the cylinder pressure wave. Andreae then concluded that the acoustic radiation is
not a result of the transmission of the cylinder pressure wave through the engine structure out to the
ambient. He further argued that the acoustic radiation is solely from the engine structure vibration, and
61
the engine structure vibration is excited by the first part (first few waves) of the pressure wave as a
forced vibration and is with the frequency of the pressure wave. After that, the cylinder gas vibration
and the structural vibration are essentially decoupled; hence the amplitudes of the two at the same
time in the subsequent development in one cycle are independent.
To illustrate this better, a power spectral density of sound radiation form the engine excited by an
impulse----hitting the engine with a hammer----was presented as in Figure 3-4. Easy to see the peak
shows up in the 4-8 KHz regime, which is closed to the frequency of the 1 't mode of the cylinder gas
oscillation. This is saying that the audible spectral power produced by the structural vibration correlates
with the spectral power of the cylinder pressure wave, and the engine structural vibration is driven by
the first few waves of the pressure wave, just as and the cylinder gas vibration.
,4030 20
10
6
4
2
0
8
12
10
16
14
18
20
Frequency (kHz)
Figure 3-4 Power spectral density of sound radiation form the engine excited by an impulse;
reprinted from Andreae [4]
The cylinder pressure is more prone to oscillation in CAl engine than in SI engine can be explained by
the fact that the heat release region in HCCI is large and more difficult to achieve pressure equilibrium,
as we mentioned in section 3.1.1.
3.1.3
The Energy Associated with CAI Engine Knock
By decomposing the waveform into a series of harmonic waves and adding up the wave intensity of
2
) associated
all of them, Eng [28] identified a quantity describing the total acoustic energy flux I (watt/m
with the pressure oscillation inside engine cylinder:
I =
c
2y
_n(_
Pn)2
P
c (A P)
2 yP
62
2
-
1 (Ap)
-yRT
2y
P
2
[3.10]
where APn stands for the amplitude of the nth harmonic wave from the Fourier decomposition. Notice
the amplitude of the resulting pressure wave is the sum of the square of the amplitudes of the individual
harmonic waves, which is not equal to the sum of the individual amplitudes as the superposition of the
harmonic waves results in both constructive and destructive interference. Easy to see that AP can be
used as a reasonable measure for the knock propensity of CAl engines, and I could be used to set up
threshold for CAI engine knock.
Noticing that the driving force of combustion chamber resonance is the rapid pressure rise due to
fast heat release, Eng [28] further considered the system analogous to a mechanical spring-mass system,
and expected the amplitude of the pressure waves directly proportional to the maximum rate of
pressure rise. Andreae [4] derived a correlation between them based on Eng's consideration:
AP
~-1 dP)
4f dt
where
f
[3.11]
-/MPRR
)max
is the oscillation frequency (f can be approximated from the
1 St
mode oscillation frequencies).
This correlation can be used to estimate the engine knock tendency with only pressure data without
resolving the acoustic waves.
3.2
CAI Engine Knock Metrics
3.2.1
Selecting the Proper Knock Metric
The engine noise can be pretty accurately recognized by human ears and easily measured with
audible devices such as microphone, as presented in Andreae's work [4]. However to consistently
measure the extent of engine knock and set up threshold for it, a more objective metric is needed.
Although audible signal can be recorded and used for analysis and judgment, it is necessary to develop a
metric from the engine combustion perspective so that improvement work regarding CAI engine knock
performance can be feasibly started from the engine combustion system, and engine knock prediction
can be feasibly conducted in simulation work.
Base on the understanding of the energy associated with pressure oscillation in engine cylinder, Eng
[28] developed a metric called knock intensity by usingPmax for P, and the corresponding Tmax
(temperature at the maximum pressure) in Equation [3.10]:
63
1 AP 2
KI =
[3.12]
.mJyRuTmax
2y Pmax
Since AP is not easily obtained without resolving the acoustic waves, Eng [28] plugged the
approximation of pressure oscillation magnitude with MPRR (Equation [3.11]) into [3.12] and achieved
the "Ringing Index":
RI
yRuTmax [ (dP
2
YPmax
[3.13]
12
dt)max
By observing the test data from his engine, Eng set up the knock threshold as RI = 2MW/m
2
.
However, test data from Lund University and MIT both have different values for RI at the setoff point of
knock [4]. This means RI doesn't serve universally for knock limit set up.
In early works, the maximum pressure rise rate in crank angle
d-J
\dO/max
was always used to set up
the knock threshold. Andreae summarized the knock criteria and measure used by researchers all over
the world [4]. The author reprinted it and added some more recent cases as below in Table 3-1.
Table 3-1
Research Group
ResearchGroup
MIT
Sandia National Lab
MIT
MIT
Lotus
Cosworth
AVL
GM
Brunel
Lund
Test Engine
Bore (mm)
80
102
80.26
87.5
80.5
87.5
86
86
76
102.6
Kncok criteria in previous research
SAE Paper
Knock Criteria
2013-01-1658
2011-01-0897
2010-01-0162
2007-01-1858
2005-01-0157
2005-01-0133
2003-01-0754
2002-01-2859
2001-01-1030
980787
KI=3MW/m 2
RI=5MW/m 2
7 MPa/ms @ 1500 rpm
MPRR 5MPa/ms
6-8 bar/"CA @ 2000rpm
6-8 bar/"CA @ 1500rpm
2 bar/CA @ 3000rpm
RI=2MW/m 2
0.5 bar oscillation 10% cycle
12 bar/"CA @ 1000rpm
Equivalent
MPRR (MPa/ms)
_
5.76
7
5
7.2-9.6
5.6-7.2
3.6
-3
7.2
It's obvious that MPRR couldn't serve universally for knock limit set up, either. As observed by Eng
[28], the MPRR value could also be very different for engine operation with different intake pressures,
even the RI value might be approximately the same at the knock limit. Also observed by Andreae via
64
comparing the audible data with calculated RI [4], the calculated RI base on MPRR data couldn't well
indicate the intensity of engine knock once MPRR is over the 5 MPa/ms threshold.
Therefore, the best knock metric is KI, as it estimates the acoustic energy flux directly from the
pressure wave amplitudes. The comparison between audible data and KI for different engine operation
conditions done by Maria further validated this point, as shown in Figure 3-5 [49].
1.0
LW
*MAPz1.
*2
4a
I br
A MAP=1. Pbar
7bwr1
0.0
00
* MAP=1.
a MAP = 1. ) b3
C
0
21
0.
.a
:E
0.1
I.
0.01
"A
*
-
1
0.1
100
10
Knock Intensity [MW/mil
Figure 3-5
Microphone signal amplitude versus knock intensity for a large data set on log-log scale;
reprinted from Maria [49]
Notice, "the data follows the trend line with slope of 0.5 (shown as solid line on the figure) for
significant values of KI (KI
1 MW/m2), which agrees with the fact that the acoustic power scales as the
microphone voltage squared. At low value of KI, the acoustic power from the pressure wave is
overwhelmed by other acoustic sources such as injection noise, block vibration, etc.", as pointed out by
Maria [49].
3.2.2
Relating the Acoustics Metric with Combustion Measurement
To address the AP and MPRR correlation problem, Maria [49] carefully and comprehensively
examined engine test data under different operation conditions. He found the ambient pressure and
fuel stratification have significant influence on the correlation. By applying the
65
1 st
law to the burning
mass inside engine cylinder, which might build up pressure wave, we have the relation between local
pressure rise rate and heat release rate:
P
where
=
[3.14]
(y - 1)4 - yP -
V
4 is the local heat release rate and V is the volume of the burning parcel. Integrating Equation
[3.14] over a short time At, we get the pressure rise magnitude:
AP = (y - 1)qAt - yP-
[3.15]
V
Easy to see the pressure rise magnitude is not only influenced by the heat release as in the
the right hand side of Equation [3.15], but also impacted by the
2nd
1st
term on
term on the right, which stands for
the work needed for expansion against the ambient pressure inside engine cylinder. Therefore, higher
intake pressure induced higher in-cylinder ambient pressure reduces the pressure oscillation amplitude
given the same amount of fuel burned. Also we can see, with smaller heat release space size, meaning
more stratified charge, the pressure rise under the same extent heat release should be smaller as well.
As mentioned before, the volumetric expansion of combustible mixture parcel is limited by acoustic
speed, so to relate the AP with MPRR, an offset term is needed to count for this effect:
AP = fl(MPRR - MPRRojjset)
[3.16]
where
MPRRoffset = yP
AV
VAt
=yP
c * S * At
VAt
= (MPRRofJset,)
ma
p
C
= yP -P
L
ft
L
[3.17]
TmTa x L
T
In Equation [3.17], S is the surface area of the parcel, L = V/S standing for the characteristic length of
the parcel, (MPRRoffset), PO, TO, LO are reference values determined by the specific combustion
conditions (intake pressure, temperature, fuel stratification extent, etc.), while Pmax and Tmax can be
66
obtained form experimental data. The correlation of AP shows great consistency with the rectified
MPRR. However, assumptions made and measurements used here need to be verified with more tests
and investigation:
1.
With the idea that pressure oscillation is caused by locally too fast heat release induced pressure
rise, there is intrinsic connection between the local MPRR and global charge MPRR, which is
used for the correlation.
2.
The amplitude AP is obtained by measuring the maximum absolute value of the pressure
amplitude of the filtered (4-20K Hz) pressure trace (as shown in Figure 2-13), while no harmonic
decomposition was depended.
3.
Corresponding reference constants are estimated experientially.
In this research, as the focus is to understand the potential of fuel stratification and EGR on CAI
engine knock mitigation, KI is used for knock measurements while MPRR is monitored to bridge up the
indicate the engine knock propensity with the combustion process. According to the comparison
between audible data and KI values, a KI = 3MW/rm2 is set as the threshold for knock in all the tests.
3.3
Controlling the CAI Engine Knock
To justify the strategy of using fuel stratification and EGR dilution for CAI knock mitigation, Prof. Wai
Cheng at MIT has conducted related theoretical analysis [51, 52]. This section reviews the analysis and
reasons the potential of fuel stratification and EGR on CAI engine knock mitigation by controlling the
heat release rate.
3.3.1
The Necessity of Extending Burn Duration
As discussed above, the driving force of CAI engine knock is the abrupt pressure rise induced by local
rapid heat release. From experiment data and related analysis we see that there is intrinsic connection
between the global and the local pressure rise rate. To reduce the knock propensity of CAI combustion,
the global abrupt pressure rise is to be relaxed as an operational goal. Below is an analysis based on
fundamental thermodynamic model and test data showing that this essentially requires extending the
burn duration of CAI combustion.
67
Starting with Equation [3.14], we consider the ideal scenario that the distribution of combustible
mixture is homogeneous (both temperature and composition) across the entire combustion chamber, so
the parameters in the equation describes the properties of the in-cylinder mixture universally. In such a
case, the second term on the right-hand-side representing the piston motion effect on pressure will be
comparatively much smaller than the rapid heat release induced pressure change as in the first term.
The volumetric heat release rate could then be used to approximate the pressure rise rate:
[3.18]
P ~ (y - 1)4
The heat release rate at crank angle 0 can be written as:
.
[3.19]
LHV( V )
V(O)
Treaction
where LHV is the lower heating value of the fuel, mj is the total mass of fuel in cylinder, V(8) is the
cylinder volume, and Treaction is the chemical reaction time scale. Combine [3.18] and [3.19] we have:
MPRR = (y - 1)
[ V(1)Treaction 1
mV LHV
[3.20]
max
Wildman [51] validated the thermodynamic model with experiment data by estimating Treaction
with 10-90% burn duration (CA90 - CA10) and V(O) with the chamber volume at CA50. The plot of
"MPRR * V(O) * -cj__9O%" versus the fuel mass mf shows as a straight line in Figure 3-6 as expected.
68
0.3
,
* 120C NVO sweep
0.25
II
*90C NVO sweep
* Temp Sweeps
o exh sweeps
0 5% dilution @90C
A
0
A 5% loan amC
A
o sto@gOC
010% dilution@120C
A10% lean@120C
*Btol@120C
C
0.2
0.15
00
0
00
0.1
10
12
14
16
Fuel mass per cycle (mg)
18
Figure 3-6 Correlation plot of MPRR to engine combustion parameters; reprinted from Wildman [51]
To increase the engine load without sabotaging the lean combustion benefits, intake boost is always
used. From engine knowledge [5] we know:
mi LHV =
NIMEP VD
[3.21]
1~
where NIMEP is the net indicate mean effective pressure,
VD
is the displacement volume, and rfj7,
is
the combustion efficiency.
Combining [3.211 and [3.20] we get:
MPRR = (y
-1)[NIMEP
r7f,i
VD
1
[3.22]
V(6*)] Treaction
where V(6*) is the cylinder volume at the MPRR point.
From Equation [3.22] we can see that MPRR increases proportionally with NIMEP if everything else
is kept constant. Yet even worse in reality: when more charge mass is stuffed into engine cylinder, the
69
auto-ignition of the mixture will be enhanced due to the shortened ignition delay, which will advance
the MPRR point and give a smaller V(O*) with the nominal MPRR point after TDC. The combustion
efficiency could be reduced with too early combustion phasing as well. The chemical reaction rate
(1/Treaction) also scales with charge density, given fixed composition and homogeneous distribution of
the charge. All the above effects associated with intake boost would significantly increase the MPRR.
This is saying the tendency of CAI engine to have higher pressure rise rate and thereby knock becomes
increasingly serious with higher engine load is required.
To get out of the dilemma, let's take a look back at Equation [3.20]: Keeping the engine load
constant or even higher, mf and LHV are not touchable. V(O) could be varied by combustion phasing,
but the effect will be very limited. (y - 1) is not a directly controlled parameter but depends on the
composition and in-cylinder temperature. Treaction seems to be the most potential variable to work
with as it can be influenced by both chemistry and charge stratification.
3.3.2
The Potential of Stratification and EGR
Both kinetics calculations and experiment data (e.g. RCM tests) indicated that the chemical heat
release process is very rapid comparing with ignition delay, and different from ignition delay, the time it
takes during the fast heat release process is not fuel sensitive. Once the major heat release process
starts, the radical degenerate branching process is so fast that there is no longer any material difference
between fuels.
In CAI engines with specified global conditions (pressure, total fuel-air ratio etc.), the actual overall
reaction time is essentially governed by the non-uniformity of the charge. With uneven temperature and
composition distribution, different mixture parcels will ignite at different time points so that the heat
release process is a sequence of very fast reactions. The incremental ignition process, i.e. the difference
of the ignition delay among the mixture parcels, governs the time scale of the whole process. This gives
stratification (temperature and composition) the opportunity to relax CAI engine combustion intensity
and abate the NVH problem.
However to use stratification as an effective strategy, the ignition delay of the mixture parcels need
to be sufficiently sensitive to the difference of temperature and/or composition of the fuel (temperature
and composition distribution within a parcel is homogeneous). To illustrate, the Livengood-Wu approach
is used [53]. To simplify the discussion, consider the charge in engine cylinder with one-dimensional
70
temperature and concentration distribution in x, equilibrium pressure P = P(t') (t' is a time point
within the period of consideration), and all the other conditions fixed across the cylinder. Then the
ignition delay of a mixture parcel can be determined as below:
[3.23]
T = T(T(x, t'), pf (x, t'), P(t'))
From the Livengood-Wu approach we know that the ignition of the mixture parcel occurs when the
integration below reaches 1:
I(,)=
[3.24]
t
dt'
to T(T (x, t'), pf (x, t'), P(t'))
The integration start time to could be as early as at Intake Valve Close. Figure 3-7 reprinted from
Prof.Cheng's report on DOE meeting can be used to facilitate the future analysis [52].
1
I (X,t+At)=
At
1=1
at
,-
(x, t)
I
I
Ax
I
I
I
I
x
X*
Next element to be burned in time
Already burned at time t
increment At
Figure 3-7 Schematic showing the ignition controlled heat release process [52]
In Figure 3-7, the portion of charge with I > 1 has already ignited by time t. For the sequential
ignition of the charge due to the temperature gradient, the next mass element Am to be ignited in the
time step At is:
71
[
[3.25]
Am = Pf(X, t') *Ax
where Ax is the increment in space dimension over which the ignition criterion will be reached in the
next time step At. Referring to Figure 3-7, Ax is given by:
[3.26]
Ax=At (t=1
Plug [3.24] into [3.26] and use that for [3.25] we have at time t and location x* (I(x*, t)
At
T(X *t)
f dt
(d
i) dpf
dpf
dx +
1):
[3.27]
pj(x*, t)
Am
=
d (j)dT
dT
dx}]
Taking the method of [3.19], we can get the local heat release rate at (x*, t):
4(x*,t)
=
[3.28]
Am
At
LHV A-
-LHV
Pf(X 0
T.xt)[Jt 0ft
0
,t'(d(1/c)dpf
t
dpf
dx
d(l/r) dT|
dT
dx)*
From Equation [3.28] we can see that stratification indeed shows potential on heat release rate control
for CAI combustion. Equation [3.28] also gives explicitly the coupling between the temperature
sensitivity of the ignition delay (d (1/T)/dT) and the temperature gradient (dT/dx), the concentration
sensitivity of the ignition delay (d (1/T)/dpf) and the concentration gradient (dpf/dx), and the
influence of this coupling to the local heat release rate. Note that the result in [3.28] depends on the
integration over the temperature and pressure history, yet because of the steep dependence of T on T,
only the higher temperature part in the late compression process is important.
From [3.28] we can also see that besides stratification and the sensitivity of ignition delay to it, the
actual value of the ignition delay of the parcel under consideration and the local fuel concentration also
72
impacts directly to the heat release process. As would be demonstrated in Chapter 4, inert gas dilution
has the effect of effectively extending ignition delay in a wide temperature regime for the fuel used. The
fuel concentration can be reduced by dilution as well. These effects justify the use of EGR for CAI knock
mitigation besides its modification to in-cylinder temperature and specific heat ratio.
3.3.3
Effectively Coordinating the Strategies
To effectively reduce the local heat release rate and control the CAI engine knock for high load
operations, special care needs to be taken when applying the proposed strategies and techniques:
1.
Sensitivity of ignition delay to fuel and temperature stratification.
Equation [3.28] shows that for stratified (temperature or concentration) in-cylinder mixture, if the
sensitivity of the ignition delay to the temperature or concentration is not correspondingly strong
enough, the heat release rate wouldn't be reduced effectively. The sensitivity of ignition delay of the
combustible mixture parcels is complex and is reflected by the fuel ignition characteristics under
different ignition conditions. Therefore to effectively use stratification for CAI engine knock control, fuel
with high sensitivity to the corresponding condition is preferred, and engine operation parameters
might need to be well designed and modified to adjust the ignition conditions in engine. This research
systematically studied the sensitivity of the ignition delay to temperature and equivalence ratio for the
fuel used as reported in Chapter 4, and the results are used to demonstrate this point together with the
engine test results.
2.
Control of fuel stratification.
Non-uniformities of fuel distribution in an engine are produced in a complex manner with dynamic
mass transfer processes involved. For fuel directly injected into cylinders to create concentration
stratification, the process of evaporation, mixing with air and residuals would all impact the status of the
mixture before ignition. With proper control, the sequential ignition might be facilitated and help on
knock mitigation, while too well mixing might not well use the stratification effect, too severe
stratification might produce local rich zones (high pf) to worsen the case, and even diesel like
compression diffusion flame combustion might occur if not enough time is available for fuel to fully
evaporate before ignition. Therefore the amount of fuel used for stratification creation and the timing of
direct injection as well as some other conditions need to be investigated and carefully designed to
generate proper stratification.
73
3.
Control of temperature stratification.
Temperature stratification naturally exists and constantly varies in engine cylinders due to the
always changing status of in-cylinder charge, the significant heat transfer to the wall, and the mixing of
the fuel, residual, and air. Because of the latent heat required to evaporate the fuel, there are also
interactions between fuel and temperature stratification. Researchers all over the world used various
methods to generate temperature stratifications, yet most of the methods are not realistic for
production engines, and the control over it is not precise, so it might not be an ideal handle for CAI
combustion control.
4.
Complex effect of EGR.
Besides the effect on ignition delay from the chemical perspective, EGR significantly impact the incylinder temperature by both the thermal mass effect and the specific heat ratio modification. Hot
residuals have also been used to enhance the auto-ignition from unstable combustion. The effect of EGR
needs to be specifically investigated and optimized for the operational goal before it's used for CAI
combustion control. This research studied the dilution effect of EGR both by RCM and engine tests.
5.
Fast heat release rate modification.
Above discussion is based on the assumption that the difference in time taken for the fast heat
release process of mixture parcels is not material, however, with fairly different temperature and
dilution conditions, the fast heat release process could be differentiated and could impact the total
combustion time. This is saying that while attention is largely given to using the inhomogeneity of
mixture for ignition delay differentiation, the reaction speed of the mixture might also be reduced to
achieve CAl knock mitigation with combustion phasing retarding, EGR dilution, or other techniques. A
demonstration of this effect is presented in Chapter 6.
6. Impact of in-cylinder pressure and other operation parameters.
From the physics of CAI engine knock we noticed that the ambient pressure inside engine cylinder
impacts the pressure rise effectively in a way by working against the thermal expansion of burned
mixture elements. With a given pressure oscillation amplitude, knock intensity, the energy associated
with the oscillation is also lower with higher ambient pressure, as seen from its derivation. Therefore
higher in-cylinder pressure could be beneficial for CAI knock mitigation with all the other conditions
fixed. Intake boosting, compression ratio modification, and some other methods could be used to
74
achieve higher in-cylinder pressure. However, when applying these techniques, special care should be
taken to make sure the affiliated impacts for on combustion phasing, fuel ignition characteristic
sensitivity to temperature and equivalence ratio, and so on, are compensated. As Dec [41]
demonstrated, due to the fuel characteristic and intake pressure differences, the fuel stratification could
have very different effect on the CAI combustion. This research systematically studied the impact of
certain engine operation parameters on the effectiveness of fuel stratification, and the interplay of EGR
and fuel stratification, as reported later in Chapter 5 and Chapter 6.
3.4
Summary
CAI engine knock is a phenomenon that abrupt in-cylinder pressure rise caused low frequency
dominated (4-8KHz) engine block resonance leads to acoustic wave radiation perceived as unacceptable
noise. The rapid in-cylinder pressure rise is primarily due to the fast heat release process by the nature
of CAI engine.
Knock Intensity, firstly derived by Eng [28], is demonstrated as a proper metric for CAI engine knock.
To relate the acoustics and combustion performance of CAl engine, Maria [49] developed a correlation
between the in-cylinder pressure oscillation amplitude and the global pressure rise rate base on a
thermodynamics model, which showed very nice consistence for a wide range of tests data.
Theoretical analysis shows that fuel stratification and EGR could be very promising approaches for
CAl engine knock mitigation, while special care needs to be taken to make sure of the effectiveness of
these strategies by understanding the fuel ignition characteristics and the in-cylinder combustion
dynamics.
75
(this page intentionally left blank)
76
Chapter 4. Fuel Ignition Characteristics and Dilution Effects
Fuel ignition behavior has significant impact on CAI combustion in engine operations [1]. To better
understand the combustion phenomenon on engine, i.e. to extract the most relevant information from
the experimental results regarding the fuel stratification and EGR effects under the specified conditions,
obtaining a comprehensive picture of the ignition characteristic for the fuel to be used is essential. In
this research, Haltermann 437 gasoline, a high octane number U.S. Federal Emission Certification Fuel
(Tier II EEE), is used for all the experiments. The specifications of the fuel can be found in appendix. This
chapter presents the results of RCM experiments that profile the ignition behavior of the fuel under
different dilution conditions corresponding to representative high load operations on CAI engine. A
simple review of fuel ignition fundamentals is also given in the first section.
4.1
Fuel Ignition Behavior and Its Measurement
4.1.1
Physics of Fuel Ignition Behavior
Auto-ignition of gaseous fuel-air mixture is set off when the rapidly increasing radical reactant
concentrations trigger an explosive acceleration of the oxidation reactions. The concentration of radical
species rises rapidlydue to chain-branching reactions and the resulting oxidation process leads to very
fast heat release rate [5].
The mechanism of chain-branching explosion has been established [5]. In chemical oxidation
processes, a large number of simultaneous interdependent reactions are involved, normally referred as
chain reactions. Among them, different reactions serve different functions: initiation reactions produces
highly reactive intermediate species or radicals from stable molecules (fuel or oxygen), following are
propagation reactions that continue the chain by forming products and other radicals from the
produced intermediates and radicals, and then the termination reactions remove the chain propagating
radicals and destruct the process. Some propagation reactions produce more reactive radicals than
consumed, and these are called chain branching reactions. When the number of radicals increases fast
enough, the reaction rate is accelerated and brings about an "explosion" of the radical molecules, and
this is called chain-branching explosion.
77
For the auto-ignition of hydrocarbon fuel, the time period after the combustible mixture is prepared
and before the rapid heat release process occurs is called ignition delay. It exists because the
hydrocarbon oxidation process involves chemical reactions comprising chain initiation, propagation, and
degenerate branching; it takes time for all of these processes to happen. By degenerate branching we
mean reactions that form unstable but long-lived intermediates, which can then either react to produce
stable molecules or active radicals, depending on the temperature. The time for degenerate branching
reactions dominates the variation of ignition delay.
Ignition delay is used to describe the ignition characteristic of a given fuel, as different fuels have
very different ignition delays under the same and different mixture conditions (e.g. temperature, charge
concentration, equivalence ratio, inert gas dilution ratio etc.). Its magnitude and variation upon different
test conditions can be used to interpret the combustion phenomenon on CAI engines.
In this research, another parameter, maximum pressure rise rate (MPRR) is also measured to reflect
the chemical kinetics characteristics of the fuel during the fast heat release process after the ignition
kicks off. It represents the speed of reaction of the tested mixture under the specific conditions.
Although theoretical analysis in chapter 3 assumed there is no material difference among fuel mixtures
in MPRR, the experiment results in this research show differently. The observations regarding MPRR will
be used to explain the engine performances later as well.
4.1.2
Measurement of Ignition Delay and MPRR
Ignition delay and MPRR data are obtained by post-process the pressure data measured from RCM
tests. The definition of Ignition delay and MPRR is illustrated in Figure 4-1 with a sample pressure trace.
Easy to see that in the first several milliseconds, the pressure rises up quickly to around 23 bar due to
the quasi-isentropic piston compression, as expected from test design calculations. After that, fuel
oxidation is set up to initiate while the pressure and temperature are kept constant inside the RCM
combustion chamber. These are the initial conditions of the mixture. A slight pressure rise can be seen
from 40 ms after the beginning of compression on. This is a two-stage heat release phenomenon. It
exists because the ignition is initiated at a relatively low temperature so that the fuel molecules would
pass through both the low and high temperature oxidation pathways [5]. The mechanism of this
phenomenon will be explained shortly. The major burst of auto-ignition occurs at around 46 ms after
compression in the figure. This gives the ignition delay defined as the time period after the completion
78
of compression and before the in-cylinder pressure reaches 50% of its maximum value. The MPRR is
also calculated base on the definition as the maximum slope during the major pressure rise process.
80
Haltermann 437 gasoline
70
0=1; dilution=0;
Tin1ua 8=32 OC ; CR=12.
60
S50
Maximum
pressure rise rate
40
30
20
Ignition delay
10 0
0
10
20
50
40
30
60
70
Time [ims]
Figure 4-1 Sample pressure trace versus time from RCM tests
4.1.3
Ignition Delay Curves
According to chemical kinetics theories, the rate of a chemical reaction is governed by temperature,
pressure, reactants concentration, and activation energy of the specific reaction [34]. For a bimolecular
reaction as shown in Equitation [4.1], the reaction rate correlates with the conditions as shown in
Equitation [4.2].
[4.1]
A+B-C+D
d[A]
P
1-EA)
oc -[A][B]expl
RT
rT_
dt
[4.2]
This format of correlation is then used to fit global reaction rates of the entire combustion process in
experiments with certain modifications on the exponents of the concentration and pressure terms for
79
the consideration of different molecule structures. Base on that, researchers fit experimental data for
ignition delay in a format as shown in Equation [4.3].
r = xP-n exp
[4.3]
)
where x, n, and y are coeeficients.
For a specific mixture (concentration and composition specified), plots made for ignition delay
against the ignition temperature (pressure condition factored into the mixture concentration and
composition) are called ignition delay curves. Figure 4-2 shows an ignition delay curve from Maria's
work [1], with ignition delay in log scale versus the reciprocal of temperature. Easy to see that there are
three regions along the curve, and Equation [4.3] doesn't work in all of them.
Temperature [K]
1000 950
900
850
800
I
I
I
1W0
High
Temp.
Region
750
I
650
700
100
Low
NTC
Region
Temp.
Regio
E
4)
E
10
10
0
1
I
LWL
U.
-
I
1.00 1.05 1.10 1.15 1.20
I
I
1.25 1.30
I
I
.
a
-
1.35 1.40 1.45 1.50
1.55
1000/Temperature [1/K]
Figure 4-2 Sample ignition delay curve; reprinted from Maria [1]
80
The dependence on temperature of ignition delay is different for initial temperatures falling in
different temperature regimes. This phenomenon is governed by the difference in hydrocarbon
oxidation pathways experienced upon the different ignition temperatures.
In the low temperature regime (below 700K in Figure 4-2), hydroperoxides (in the form of ROOH,
where R is an organic radical formed by abstracting a hydrogen atom from a fuel hydrocarbon molecule
[5]) are produced through the chain propagation process as the most important intermediate for
degenerating branching. With the initial temperature increasing, the production rate of ROOH increases
as well, giving an exponentially decreasing ignition delay as described in Arrhenius form equation for
reaction rate.
In high temperature regime (greater than 900 K), hydrogen peroxides (H2 0 2 ) formed from chain
propagation reactions dominates the degenerating branching process for producing more radicals as it
loses its stability and decomposes into OH radicals when temperature is higher than 500 'C. The
production rate of radicals increases and gives an Arrhenius format decreasing of ignition delay with the
initial temperature going up, similar to the low temperature regime situation.
In the intermediate temperature regime (700K~900K), the production of ROOH gradually terminates
due to the decomposition of R0
2
radical (reverse reaction takes over the forward synthesis with
temperature going up), which is from chain initiation reaction products and is to be used as the
reactants for ROOH production. The oxidation pathway dominated by H20 2 has yet to be opened up, as
the hydrogen peroxide molecule is relatively stable at such temperatures. Therefore, with the initial
temperature increasing, the production of radicals are not ramming up as it is in the higher or lower
temperature region, and therefore ignition delay doesn't shorten either. This creates a so-called
negative temperature coefficient (NTC) regime. In this region, the sensitivity of ignition delay to the
change of temperature goes to minimum comparing with in other temperature regimes.
The diagram in Figure 4-3 organized by Hahn [54] and edited by the author well presents the above
discussion.
81
Chain Initiation
Chain Reaction mechanism
chain
Propagation
RH+ 02 -+ i+
Low Temperature
hIo
NTC regime
High
0 2 t4RO
t
2
I
RU2-
RH+H
roducton in
equilibrium
iOOH
QOOH + 02 - OOQOOH
H-+R
eranchng a
R2O
- =R0 +
H: or starts to
* decompose.
H
2+H0 2+M
perature
2
-
-+
H 20
2
H20 2 + 02 +M
Brancing ent hyen peroxde)
>
±:::0
M
Characters of Hydrocarbon Chemical Oxidations, where RH, R, and QOOH
boy
are hydrocarbon species, 02 and OH are oxidizer, and M is 3 rd1rbody.
O-H + O
j
Degenerate
Branching
(radical
production part)
Figure 4-3 Chain reactions and NTC region mechanism; reprinted from Hahn [541
4.2
Ignition with Fixed Charge Density
Charge density is the amount of fuel-air mixture in unit volume, measured in "Kmol/m 3" in this
research. In engine operations, the compressed charge density in the engine cylinder is determined by
engine geometry and intake conditions. To understand how fuel stratification, i.e. the equivalence ratio
distribution, impact the heat release processes, tests with constant charge density but different
equivalence ratios are conducted on RCM.
The charge density chosen for this test sets is 0.75 Kmol/m
3
to give an initial temperature of 800K at
an initial pressure of 50 bar, which is a typical after-compression condition for high load CAI operation.
The ignition temperature is swept from 600 K to 1100 K to give a full picture of the ignition delay curves.
The equivalence ratio is varied from 0.3 to 0.7 with an interval of 0.2 to simulate the lean burn condition
of CAI engines. The results are shown in Figure 4-4.
82
Temperature (*K)
1100
Halterman
437
gasoline
1000
900
600
700
800
0.75 kmol/m 3
(P=50 bar @T=8000K)
100
No dilution
*10
0
1
0.9
1.0
High T region
1.1
1.2
1.3
1000/T (*K1)
- Sensitive to T
<D=0.3
0=0.5
1.5
NTC region
tdelay
- Less sensitive to 4
1.4
13
0
A
Idelay
- Sensitive to (D
Less Sensitive to T
0=0.7
1.6
Low T region
"Cdelay
* Less Sensitive to (F
- Sensitive to T
Figure 4-4 Ignition delay with fixed charge density 0.75 KmoI/m
3
It is clear that in different temperature regimes, ignition delay has different dependence on
temperature and equivalence ratio. For cases with initial temperature lower than 650 K, ignition delay
shortens nearly exponentially with temperature for all three fuel-air mixtures, but there is no significant
change from the variation of equivalence ratio. So in the low temperature regime, the ignition delay is
sensitive to temperature variation, but not to equivalence ratio differences.
For cases with initial temperature over 1000K, the ignition delay for all the mixtures decreases in a
pace governed by Arrhenius equation as well, and the dependence on equivalence ratio is relatively
small. Thus in the high temperature region, the ignition delay is also sensitive to temperature variation,
but not to equivalence ratio differences.
Due to the limit of measurement capability, tests with too high initial temperature are not possible
for high equivalence ratio mixtures that might give ignition delay shorter than ims. However, shock tube
experiments on self-ignition characteristics of SI engine model fuels (iso-octane, methanol, iso-octane &
n-heptane blends etc.) conducted by K. Fieweger and his colleagues in the Institute of General
83
Mechanics, Aachen, Germany could help on validating the above hypothesis [55]. Shown in Figure 4-5
are ignition delay curves obtained from iso-octane mixtures with fixed total charge density determined
by the initial pressure (here it's 40 bar). It is clear that ignition delay curves do converge in the very high
temperature regime, showing a low sensitivity to equivalence ratio differences. Notice in this graph, rj is
an "ignition delay" defined as the time period between the moment when after-compression conditions
are reached and the moment the first flame kernel appears. This is brought about by the inhomogeneity
(hot spots) inside the shock tube combustion chamber and could be measured by both shadowgraphs
and emission signals. This "ignition delay" doesn't represent the chemical kinetic characteristics of the
fuel, but reflects the beginning of the inhomogeneous, deflagrative phase. On the other hand, T - DDT,
the ignition delay between the start time point and the instance of explosion (first dramatic pressure
rise), stands for the homogeneous self-ignition of a larger volume of the test gas. Therefore the
interested data to this research is -r- DDT that is interpreted as the best approximation for the
chemical ignition delay time.
100!
1300 1200 1100
1000
900
T5
800
700
iso-Octane
p =40 ±2 bar
10
......---....
-
-
-.--
-
[ins]
~
-
-
---
=2.0
x-----0=2.0
___---
A 0=1.0
A0=1.0
+ =0.5
.7
0.8
0.9
1
1.1
1.2
--
-----
-
0
0
0.1
-
1.3
-rDDT
;
T-DDT
-
'C1
T-DDT
1.4
1000 K/ T5
1.5
Figure 4-5 Ignition delay of iso-octane in high temperature regime; reprinted from K. Fieweger [55]
For cases with initial temperature between 650K and 950K, the NTC behavior shows up on all three
curves. Notice in this regime, although the slopes of the three ignition curves look very similar, the start
points and end points of NTC region (inflection points in the two ends of the regime) are different for
84
different equivalence ratios. With higher equivalence ratio, the NTC region is smaller----later start and
earlier end of the NTC behavior.
With this knowledge, we believe in that by restraining the in-cylinder conditions in the NTC regime,
the higher sensitivity of ignition delay to equivalence ratio of the fuel could help enhance the sequential
ignition of the stratified charge and thereby weaken the engine knock tendency. On the other hand, if
fuel-air mixture falls into the equivalence ratio insensitive regime (too high or too low temperature
regime), fuel stratification strategy might not work effectively for sequential burn creation.
Ignition with Fixed Fuel Density
4.3
Fuel density is the amount of fuel contained in unit volume, indicated by the total fuel-air mixture
charge density under stoichiometric equivalence ratio, and dilution condition; the unit is "Kmol/m
3
.
Since all the operating conditions are lean, by fixing the fuel density, the maximum energy output in unit
volume is fixed and the effect of different levels of air dilution (simulating boosting on engines) and inert
gas dilution (simulating EGR on engines) can be examined and compared on an equal foundation. This
approach isolates the impact from other factors and highlights the dilution effects, so that the behavior
of CAl engine operated with intake boosting and EGR can be better interpreted with this information.
The tests are conducted with three different dilution set ups: 1) using nitrogen as inert diluent with
a fuel density 0.4 Kmol/m 3 ; 2) using carbon dioxide, water vapor, and air as diluent (simulated EGR) with
a fuel density 0.4 Kmol/m 3 ; 3) using simulated EGR diluents with a fuel density 0.6 Kmol/m
4.3.1
3
.
N 2 as Diluent
N2 is chosen as the inert dilution gas for the first set of experiments based on two considerations: 1)
simple inert diluent composition with only N2 could isolate the possible chemical effects from CO 2 and
H20; 2) nitrogen takes the bulk portion in the real EGR gas for lean combustion CAI engines (about
73.4%~78.4% for equivalence ratio varying from 1 to 0.1), therefore it could best represent the property
of EGR diluents than other gas.
The effects of both air dilution and inert gas dilution are tested orthogonally with a matrix as shown
in Table 4-1. In the table are the after-compression charge density values under specified dilution
conditions. To simulate the high load operations of CAI combustion with typical engine geometry and
85
intake conditions, the fuel density is kept constant at 7.825 mol/m 3 , which is determined from the
charge density 0.4 KmoI/m 3 at stoichiometric equivalence ratio and with no inert gas dilution. By varying
the equivalence ratio, different levels of air dilution effects were tested, and by varying the dilution ratio,
different levels of EGR effects were examined. The dilution rate is set upon considerations over the
realistic EGR rate of a CAI engine, and the equivalence ratios are set based on the purpose of high load
operation simulation and are restricted by the charge capacity limit.
For each point in the matrix, temperature after compression was swept in the range of 600K to 1100
K. This was realized by adjusting the RCM compression ratio and the temperature of RCM combustion
chamber walls.
Table 4-1
RCM Run Matrix----N 2 as Diluent
Equivalence
Ratio
1.
Dilution Rate (by mole fraction)
0
0.2
0.4
1
0.400 (Kmol/m )
0.500
0.667
0.8
0.498
0.623
0.830
0.6
0.661
0.827
1.102
3
Air Dilution Effect
The ignition delay curves demonstrating equivalence ratio dependence with different inert gas
dilution rates are shown in Figure 4-6, Figure 4-7, and Figure 4-8.
86
950
900
Temperature [K)
800
750
850
700
650
Dilution=0-
100
100
E
Phi=1.0
0
10
10
0
0
M)
Phi=0.6
1
1.1
1.2
1.3
1.4
1.5
1000/Temperature [1/K]
Figure 4-6
Ignition delay----air dilution effect with inert gas dilution rate=O
950
900
850
Temperature [K)
800
750
700
-
650
I
100
100
Phi=1.0
Phi 0 6
Cu
15
0
10
10
0)
Phi=0.8
I
.11.2
1.3
1.4
1.5
1
1000/Temperature [I/K]
Figure 4-7 Ignition delay----air dilution effect with inert gas dilution rate=0.2
87
Temperature [K]
950
900
800
I
S
100 :
850
750
700
I
650
I
Dilution=0.4
Phi=0.8
-
1
Phi=1.0
0
10
Phi=0.6
1
I
1.1
1.2
I
1.3
1
1.4
1.5
1 000/Temperature [1/K]
Figure 4-8
Ignition delay----air dilution effect with inert gas dilution rate=0.4
From the pictures we see that with the same amount of inert gas dilution, the ignition delay curves
generally don't vary from each other significantly upon different levels of air dilution, although the
departure of the curves to each other in different temperature regimes is not all the same.
In the low temperature regime (<750K), overlapping ignition delay curves are seen for mixtures with
different air dilution rates, regardless the level of inert gas dilution. This is saying for combustible charge
with the same fuel density and relatively low initial temperature, the ignition delay of the tested fuel is
not sensitive to equivalence ratio.
However for tests with ignition temperature from 750 K upwards, differently blended fuel-air
mixture start showing slightly different NTC behaviors. With more air dilution, the ignition delay is
shortened in general. This could be resulted from higher radicals generation rate with more available 02
for both low temperature and high temperature oxidation pathways. This observation is further
confirmed by comparing the N2 diluted cases and the simulated EGR cases. Notice the very different
NTC behavior of the ignition delay curve with no inert gas dilution and stoichiometric equivalence ratio
in Figure 4-6, which shows that without inert gas dilution, more available oxygen could cause shortened
ignition delay for tests in the NTC regime quite effectively.
88
In the even higher temperature regime, certain equivalence ratio dependences are seen in higher
inert-gas-diluted cases. This is saying that for engine operations with the same fuel input, highly boosted
intake could shorten the ignition delay of fuel mixture by both increasing the in-cylinder temperature
and shortening the ignition delay with more oxygen, so that the auto-ignition would be enhanced and
higher propensity to knock could be induced.
In summary, adding more air for the same amount of fuel doesn't significantly impact the ignition
behavior in a significant manner in the tested temperature regime. The effect of intake boost is more
from a physical than chemical perspective. However we should notice under relatively high ignition
temperature conditions, more available air could slightly shorten the ignition delay of combustible
mixtures.
The MPRR curves with corresponding conditions to the ignition delay curves above are shown in
Figure 4-9, Figure 4-10, and Figure 4-11. Clearly MPRR curves show very different dependence to
temperature in the low and high temperature regimes separated at around 700K. In the low
temperature regime, MPRR drops down drastically with the decrement of initial temperature.
In the
high temperature regime, it barely changes with the variation of initial temperature. With higher inert
gas dilution, the temperature dependence of MPRR in the high temperature regime gets slightly
enhanced. It's obvious that equivalence ratio doesn't make much impact on MPRR in the entire tested
temperature range, as all the curves are mostly entangled with each other, although slight differences
can be observed for curves with different equivalence ratios in the near NTC regime.
Therefore for the same amount of fuel in unit volume, the heat release process can be effectively
slowed down if the ignition temperature can be directed into a lower temperature regime, while simply
diluting the mixture with more air might not help. Implications on CAI engine operations are that
although the actual heat release process is of much smaller order of magnitude time period length
comparing with the ignition delay, the chemical kinetics effects could still kick in to help when it comes
to an extend that the fuel combustion happens under a relatively low temperature. This could be the
mechanism behind combustion phasing adjustment for CAI knock performance variation.
Notice in Figure 4-11, the highly diluted case with equivalence ratio 0.6 has a different trend as
other curves. This is saying with very lean mixture, the speed of reaction could actually get slowed down,
as the optimized ratio for fuel and oxygen elements to collide and interact is passed far away. This
89
implicates that low equivalence ratio mixture has the potential to contribute on burn duration extending
on CAI engines.
Temperature [K]
950
1Inn
100
650
700
750
800
850
900
Dilution=0
100
Phi=0.8
E
10
10
Phi=0.6
CL
CIr
Phi=1.0
2:
1
1.5
1.4
1.3
1.2
1.1
1 000/Temperature [1 /K]
Figure 4-9 MPRR---- air dilution effect with inert gas dilution rate=O
Temperature [K]
950
100
900
850
D(iluion=-0.2
E
a
a.:
650
700
750
800
100
Phi=0.8
10
10
-
Phi=0.6
I
1Phi=1.
1
1.4
1.5
1
.11.2
1.3
1000/Temperature [1/K]
Figure 4-10 MPRR---- air dilution effect with inert gas dilution rate=O.2
90
Temperature [K]
950
100
900
800
850
750
700
650
100
Dilution=0.4
E
E
10 -10L
0.
a.
Phi==.6
1
1
1.1
:
1.2
i
1.3
i
i1
1.4
1.5
1000/Temperature [1 /K]
Figure 4-11 MPRR---- air dilution effect with inert gas dilution rate=0.4
Therefore MPRR is a parameter depending on ignition temperature, especially in the lower
temperature regime, and this could be one of the important factors that make combustion phasing
impact the burn duration on CAI engines besides sequential ignition. In general, air dilution doesn't vary
the fast heat release much, however with very lean mixture, the reaction could be slowed down possibly
due to the departure of fuel-air ratio from stoichiometry.
2.
Inert Gas Dilution Effect
The ignition delay curves demonstrating equivalence ratio dependence with different inert gas
dilution rates are shown in Figure 4-12, Figure 4-13, and Figure 4-14. Comparing with the air
dilution
results, we can see that the inert gas dilution effectively extends the ignition delay for mixtures
with
different equivalence ratios in all the temperature regimes. This means by having higher EGR rate,
CAI
engines could have a later combustion phase, and therefore a better anti-knock performance
as
reported in literatures [37]. Although more oxygen will also be induced into engine cylinders with inert
exhausts gas for lean combustion, the effect of inert gas still dominates. Notice that inert gas dilution
shows higher impact in higher temperature regimes although the change is presented in a log-scale
coordinates. This means the effectiveness of EGR dilution on combustion phasing modification does
91
have dependence on the in-cylinder conditions, and optimization based on the characteristics of the fuel
would be needed for the best outcome.
Temperature [KJ
950
C
0
900
800
850
650
700
750
100
A100
10
10a
C
.2
C
Diluon=0
Dilution=0.2
Diluion=0.4
1
1.1
1.5
1.4
1.3
1.2
1000/Temperature [1/Kj
Figure 4-12 Ignition Delay---- inert gas dilution effect with equivalence ratio=1.0
Temperature [KI
950
900
850
650
700
750
800
100
100
cc
CU
10
10
.2
10
C
0
CM
Diluton=P
Dilution=07.2]
Dilution=0.4
1
1.1
1.1
I
1.2
I
1.3
1I
1.4
'
I
1.5
1000/Temperature [1/K]
Figure 4-13 Ignition Delay ---- inert gas dilution effect with equivalence ratio=0.8
92
900
950
100
Temperature [K]
800
750
850
650
700
.. Phi=O.6j
100
U,
E
4)
a
cc
10
C
0
C
0,
10
C
0
0
CM
Dilutlon=0
Dilution=0.2
Dilution=0.4
1
1.1
1.2
1.3
1.4
1.5
1000/Temperature [1/K]
Figure 4-14 Ignition Delay ---- inert gas dilution effect with equivalence ratio=1.0
The MPRR curves of the mixtures under corresponding conditions as above are shown in Figure 4-15,
Figure 4-16, and Figure 4-17.
Temperature [K]
950
900
850
800
750
700
650
100
Phi=1.0
E)
M
CL
E
Q
10
oO
10 a-
Dilution=0
CkL
Diuin=0.2
a.
Dilution=0.4-
1
1.1
1.2
1.3
1.4
1.5
1000/Temperature [1 /K]
Figure 4-15 MPRR---- inert gas dilution effect with equivalence ratio=1.0
93
Temperature [K]
950
100
900
650
700
750
800
850
100
Phi=0.8
0..
a
10
10
Dilution=0.4
01
0~
Dilution=-0.2
Dflution=0
I
1.5
1.4
1.3
1.2
1.1
1000/Temperature [1 /K]
Figure 4-16 MPRR---- inert gas dilution effect with equivalence ratio=0.8
Temperature [K]
950
100w
900
650
700
750
800
850
100
-Phi=0.6
.
10
10
a02
Dilution=0.4
.
Dilution=0.2
cc
Dilution=O
I
1.1
1.2
1.4
1.3
1.5
1000/Temperature [1/K]
Figure 4-17 MPRR----boost effect with equivalence ratio=0.6
94
It is clear that inert gas dilution helps to slow down the heat release process as well in the
temperature range tested, although the effect is much stronger in the temperature regime of 675 K to
850 K (-NTC region) than the too low or too high temperature regimes. However it's worth to notice
that the dilution effect is not "linear", i.e. comparing the change from dilution rate 0 to 0.2 with 0.2 to
0.4 we can see that higher inert gas dilution rate makes more impact. This is saying that on CAl engines,
the fuel combustion process can be more effectively impacted by inert gas dilution when the in-cylinder
conditions are adjusted into the NTC regime, and the impact is stronger with higher dilution rates.
3.
Air Dilution vs. Inert Gas Dilution
To understand the relative importance of the two dilution techniques in a more quantitative manner,
the ignition delay data is fitted based on a comparison between diluted cases and the base case with
stoichiometric equivalence ratio and no inert gas dilution. The format of the data fitting is inspired by
Wooldridge isooctane results [56] and presented as below:
[4.4]
T = r(Xd = 0)(1 - Xd)acpb
where the
Xd
is the mole fraction of inert dilution gas. To highlight the relative importance between air
and inert gas dilution, the complication of temperature dependence is avoided by setting the
coefficients as constants. The result is shown in Figure 4-18.
Temperature [K]
900
850
800
700
750
650
a =-1.5
b =0.3
N2 dilution
A
Dilution
0.0 0.2 0.4
44
S
1.0
0.8
0.6
1.1
1.2
1.3
1.4
A
v
*
A
4
*
O
1.5
1000/(T[K])
Figure 4-18 Air dilution vs. inert gas dilution.
95
With the minimum departure of the tested data to the fitted equation by a R2 of 0.985, coefficients
are achieved as a=-1.5 and b=0.3, meaning inert gas dilution extends the ignition dilution in a 5 times
higher power than air dilution's effect on shortening the ignition delay. This means on CAI engines, cool
EGR dilution could be used as a very effective lever for combustion phasing control at high load
operations, as the possible combustion phase advancement caused by intake boost would be limited
from the chemical perspective comparing with EGR.
To see clearly how inert gas dilution prolongs the ignition delay, a plot for the change made by it
(T/T (Xd =
0)) is plot versus dilution rate in Figure 4-19 as below. For a twice as long ignition delay
comparing with the base case, around 37% inert gas dilution is required, while another 33% more
dilution gives a six times as long ignition delay. This explains the "nonlinearity" of inert gas dilution
effects: a high dilution rate is needed to make it effective. This also implicates that dilution gradient
would produce significant stratification of ignition at high dilution rates to strengthen the sequential
ignition inside engine cylinders.
6
T C1
Xd)
3
2
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
Dilution (xd)
Figure 4-19 Dilution rate dependence of inert gas dilution effectiveness
4.3.2
Simulated Combustion Products as Diluent
Carbon dioxide (C0 2 ) and water vapor (H20) are used as diluents for a new set of RCM tests to
achieve incremental knowledge of inert gas dilution effects to ignition delay of the specific fuel. By
comparing these results to the previous test set, a closer understanding on the actual impact of EGR on
96
CAI combustion could be achieved. Two subsets of tests were conducted with charge densities of 0.4
Kmol/m 3 and 0.6 KmoI/m 3 separately for the purpose of understanding the fuel density dependence.
1.
Simulated EGR vs. Inert Gas Dilution
Figure 4-20 to Figure 4-22 show the comparison of ignition delay data for the simulated EGR cases
and inert gas dilution cases with the base case charge density 0.4 Kmol/m 3 . Since cases with no inert gas
dilution are the same for the two sets experiments, we only compare the results with dilution rate 0.2
and 0.4. We can see that the general trend of ignition delay against temperature of the tests using
nitrogen as inert gas diluent has excellent consistency with tests using simulated combustion products
as inert gas diluent. However for data points with temperature higher than 750K, simulated EGR tests
give a slightly shorter ignition delay. This is because the actual oxygen fraction is higher for these tests,
and correspondingly the inert gas fraction is lower. The fuel/oxygen equivalence ratios for simulated
EGR tests in Figure 4-21 (nominal Phi=0.8) are 0.76 and 0.71 for "dilution rate" 0.2 and 0.4 respectively,
and this makes the actual fraction of inert gas diluent less than the nominal ones: they are 0.17 and 0.33
respectively. This difference is magnified by higher oxygen concentration in the lower equivalence ratio
cases.
900
850
800
750
700
650
100-I
0
100
0
0
P 10
C
10-
0
0
10
0
0)
0
"
,N2
0
0
.
Dilution 0.2
N2 Dilution 0.4
EGR Dilution 0.2
EGR Dilution 0.4 .
I
1.1
1.2
1.3
1.4
1.5
1000/Temperature [I/K]
Figure 4-20 Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=1.0
97
650
700
750
800
850
900
Phi=O.8
100-
+-
100
0
W,
E
A
*xA
C
0
10.
A
-
10
A
A
N2 Dilution 0.2
o N2 Dilution 0.4
EGR Dilution 0.2
A
A
EGR Dilution 0.4
I .1
1.1
1.4
1.3
1.2
1.5
1000/Temperature [1/K]
Figure 4-21 Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=0.8
950
100
900
800
850
650
700
750
Phi=O.6
100
E
I~
10-
1:t*
*
10
1~1~
1~z
* **
**
*
N2 Dilution 0.2
N2 Dilution 0.4
EGR Dilution 0.2
EGR Dilution 0.4
I
i
1.1
1.2
1.3
1.4
1
1.5
1000/Temperature [1/K]
Figure 4-22 Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=0.6
The comparisons of MPRR are shown in Figure 4-23, Figure 4-24, and Figure 4-25. Easy to see that
the MPRR for the two-dilution test sets overlaps pretty well, so the minor change of oxygen and inert
98
gas fractions doesn't impact much on the major heat release process. Similar to previous tests, the
dilution effect is seen more obviously in the NTC regime than the low and high temperature regions.
950
m
1W
900
850
800
750
700
650
100
Phi=1 .0
000
00
W)
E
0-
10-
10
2
n
.
0
o N2 Dilution 0.2
*
*
*
N2 Dilution 0.4
EGR Dilution 0.2
EGR Dilution 0.4
0
1
1
1.1
1.2
1.3
1.4
1.5
1.6
1000/Temperature [1/K]
Figure 4-23 MPRR-Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=1.0
100
950
900
800
850
750
700
650
100
Phi=0.8
ZA
A+
AA
E
10
0.
10
i
0
A
+
N2 Dilution 0.2
N2 Dilution 0.4
EGR Dilution 0.2
EGR Dilution 0.4
A
A
1.1
1.2
1.3
1.4
1.5
1
I .6
1000/Temperature [1/K]
Figure 4-24 MPRR-Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=0.8
99
10(
850
900
950
800
650
700
750
100
Phi=0
1
o
10
*-
l
a.
*
N2 Dilution 0.2
N2 Dilution 0.4
EGR Dilution 0.2
EGR Dilution 0.4
1.1
1.2
*
1.4
1.3
1.5
1.6
1000/Temperature [1/1(
Figure 4-25 MPRR-Dilution Effect of N2 vs. Simulated EGR with equivalence ratio=0.6
2.
Fuel Density Impact
The impact of fuel density to ignition delay is also tested to see the relative significance of engine
gas
load increment to EGR dilution, as shown in Figure 4-26, Figure 4-27, and Figure 4-28. The inert
trend of
diluent in this set of tests is also simulated combustion products. It is obvious that the general
extends
dilution effects in higher fuel density tests is similar to lower fuel density tests: inert gas dilution
ignition delay in general, and higher dilution rate enhances the effect.
It's also obvious that increment of fuel density shortens the ignition delay of the mixture
in the
substantially. Notice this effect also has temperature dependence. The change of ignition delay
on CAI
higher temperature regime is larger than that in the lower temperature regime. This means
free
engines, having more fuel in an engine cycle requires a higher dilution rate to maintain the knock
or higher intake
performance, especially for higher ignition temperature scenarios (e.g. high CR
gas dilution
temperature). In other words, on the "ignition delay vs. ignition temperature" map, inert
results in upwards movement of ignition delay curves, and increasing charge density causes downwards
movement of ignition delay curves. These movements all have larger magnitude in the relatively higher
would cause
temperature regime than in the lower temperature regime. However, too high EGR rate
misfire. Therefore a subtle adjustment over power output and EGR rate is need to keep the engine
running free of both knock and misfire.
100
900
100
850
0.6kmol/m3
o 0.6kmol/m3
-4- 0.6kmol/m3
- 0.4kmol/m3
-*0.4kmol/m3
-
E
750
700
100
Dilution 0
Dilution 0.2
Dilution 0.4
Dilution 0
Dilution 0.2
0.4kmol/m3 Dilution 0.4
-A.2
800
10
10
P=1
Phi=1 .0
1.1
1.2
1.5
1.4
1.3
1000/Temperature [1/K]
Figure 4-26 Ignition Delay----Fuel density Impact with equivalence ratio=1.0
900
1007
e
-v
E
0
-0-
850
0.6kmol/m3
0.6kmol/m3
0.6kmol/m3
0.4kmol/m3
0.4kmol/m3
0.4kmol/m3
750
800
650
700
Dilution 0
Dilution 0.2
Dilution 0.4
Dilution 0
Dilution 0.2
Dilution 0.4
100
10-
10
e
e
Phi=0.8
-
1-
1.1
1.2
1.3
1.4
1.5
1000/Temperature [1/K]
Figure 4-27 Ignition Delay----Fuel density Impact with equivalence ratio=O.8
101
650
700
750
800
850
900
0.6kmol/m3 Dilution 0
0.6kmol/m3 Dilution 0.2
+4-0.4kmol/m3 Dilution 0
*
0.4kmol/m3 Dilution 0.2
0
0.4kmol/m3 Dilution 0.4
4
100-
100
*
0
10-
C
10
*
.0
4.
* ~*4~*~4~
0,~
4
*
Phi=0.6
1-
1.4
1.3
1.2
1.1
1.5
1000/Temperature [1/K]
Figure 4-28 Ignition Delay----Fuel density Impact with equivalence ratio=0.6
The inert gas dilution effects to MPRR are compared for different fuel density cases, as shown in
Figure 4-29, Figure 4-30, and Figure 4-31.
900
800
850
650
700
750
+nn
2'10
100
10
W.
*
---
1x
+
A,
1.1
0.6kmol/m3 Dilution 0
0.6kmol/m3 Dilution 0.2
A6krnm/m3 Dilution 042
0
0.4kmol/m3 Dilution 0
0.4kmol/m3 Dilution 0.2
0.4kmol/m3 Dilution 0.4
1.2
1.4
1.3
1.5
1000/Temperature [1/K]
Figure 4-29 MPRR----Fuel density Impact with equivalence ratio=1.0
102
950
900
850
800
750
700
100 100
eD -
650
100
100
-- --
.
0
-
N0
10
10
0.
0.6kmol/m3 Dilution 0
0.6kmol/m3 Dilution 0.2
0.6kmol/m3 Dilution 0.4
-v- 0.4kmol/m3 Dilution 0
+ 0.4kmol/m3 Dilution 0.2
-- 0.4kmol/m3 Dilution 0.4
1.1
1.2
V
P
1
I
1.3
1.4
1.5
1000/Temperature [1/K]
Figure 4-30 MPRR----Fuel density Impact with equivalence ratio=0.8
I
950
UU
900
850
800
750
700
650
100
-
CL
*
10a.:
c1
10
4
0.6kmnol/mn3
Dilution 0
0
0.6kmno/m3 Dilution 0.2
*
0.4kmnol/m3 Dilution 0
0.4kmnol/mn3 Dilution 0.2
+0.4kmnol/mn3 Dilution 0.4
---
Phi=O.6
1
1
1.1
1.2
1.3
1.4
1.5
1000/Temperature [1/K]
Figure 4-31 MPRR----Fuel density Impact with equivalence ratio=0.6
Similar to lower fuel density tests, in the higher fuel density tests: 1) Inert gas dilution relaxes the
major heat release process; 2) the effect is more obvious in the NTC region; 3) higher dilution rate is
103
more impactful. It's also obvious that higher fuel density accelerates the heat release process in the
entire temperature range to give a higher MPRR. This could be simply because more energy is released
in unit volume during specific time period for tests with higher fuel density. Therefore although inert gas
dilution can fuel stratification can be used for combustion phasing rectification and burn duration
extending, more fuel in the cylinder would directly brings about higher driving force for engine knock.
4.4
Summary
A series of RCM experiments were conducted to comprehensively understand the ignition
characteristics of the fuel used in this research. It was found:
1.
The sensitivity of ignition delay to temperature and equivalence ratio varies with the initial
conditions. In this research, the fuel used shows NTC behavior in the temperature regime of
650K to 950 K for mixtures with fixed charge density at 0.75 Kmol/m
3
and equivalence ratio 0.3,
0.5, and 0.7 respectively, where the differences of ignition delay is maximized by varying
equivalence ratio. With lower or higher temperature, the sensitivity of ignition delay to
temperature is higher but to equivalence ratio is reduced. This means to make fuel stratification
effective on extending the CAI engine heat release process, the in-cylinder conditions need to be
adjusted into certain temperature regime with substantial sensitivity of ignition to equivalence
ratio.
2.
Air dilution doesn't vary the ignition behavior much but moderately shortens the ignition delay
in higher temperature regime. Inert gas dilution extends the ignition dilution in a much higher
power than air dilution's effect on shortening the ignition delay, and the effect is also stronger
in higher temperature regime. The inert gas dilution effect becomes stronger with higher
dilution rate.
3.
MPRR is a parameter depending on ignition temperature, especially in the lower temperature
regime, and this could be one of the important factors that make combustion phasing impact
the burn duration on CAI engines besides sequential ignition.
In general, air dilution doesn't
vary the fast heat release much, however with very lean mixture, the reaction could be slowed
down possibly due to the departure of fuel-air ratio from stoichiometry. In the other hand, inert
gas dilution effectively reduces the MPRR in the temperature regime of 675 K to 850 K.
104
4.
Tests with simulated EGR gas as diluent and with nitrogen as diluent give the same conclusion
except the effect on shortening the ignition delay is weakened in higher temperature regime
due to the higher oxygen fraction in the simulated EGR gas.
5.
Higher fuel density shortens the ignition delay, especially in higher temperature regime; higher
fuel density increases the MPRR, especially in NTC region.
105
(this page intentionally left blank)
106
Chapter 5. Boosted CAI Engine with Fuel Stratification and EGR
Fuel stratification means uneven distribution of the fuel concentration inside engine cylinders
before ignition. As described in Chapter 3, fuel stratification has the potential of enhancing sequential
combustion and therefore mitigating the knock propensity on CAI engines. While stratification always
exists to some extent inside engine cylinders, the extent of stratification depends on the fuel-air mixing
process and can be impacted by factors including fuel delivery methodology (PFI vs. DI), DI injection
timing, cylinder and combustion chamber geometry, valve timing, intake conditions, engine speed (time
available for mixing and turbulence intensity), and so forth. In this research, in-cylinder fuel stratification
was adjusted via the variation of the DI injection timing (start of injection) while most of other
conditions were kept constant for comparisons. For cases with both PFI and DI injection, the split ratio
of the fuel amount delivered via the two portals was also varied to create different levels of fuel
stratifications.
EGR is also an effective tool for reducing CAI engine knock tendency. However, clear understanding
of the complex effects of EGR on CAl combustion still requires further investigation. In this research, EGR
rate was swept by throttling the EGR valve for differently configured CAI combustion operations with
exhaust pressure higher than the boosted intake pressure and fuel stratification.
This chapter presents the knock and combustion performance of the CAI engine with fuel
stratification and EGR under intake boosted conditions.
Responses to Stratification at Compression Ratio = 19
5.1
As described in Chapter 2, the test engine is modified from a Renault diesel engine with an original
compression ratio of 19. This is to take advantage of the enhanced start up ability, low load operation
range, and high thermal efficiency of the higher CR configuration. This also allows relatively high octane
number fuels be used on this engine for better knock performance without deteriorating the ignition
performance. The first batch of experiments was conducted with this set up.
5.1.1
Map of Operation Range for the CR19 Engine
107
To define the domain of operation, the low load limit and high load limit of the engine under
different EGR rates and intake pressures. The intake temperature is fixed at 80 'C throughout the entire
research except for tests focusing on intake temperature effect studies. The value is set based on
considerations of water vapor condensation in EGR and the ability of initiating CAI operation. The engine
speed is fixed at 1200rpm for all the tests except for engine speed effect study. The exhaust pressure is
throttled to 0.1 bar higher than the intake pressure to make sure the accessibility of EGR.
As shown in Figure 5-1, intake pressures of 1.2 bar, 1.5 bar, and 1.9 bar were separately tested. For
each intake pressure condition, the EGR rate is swept from 0 to the upper limit (determined by mixture
equivalence ratio which is less than 1). For each EGR rate condition under the specific intake pressure,
the amount of fuel injected is varied to probe the load limits. The fuel is delivered via PFI portal with a
fixed timing (70 *CA BTDC intake) to eliminate the possible fuel stratification effects (valve wetting). The
low load limit is set by misfire and the high load limit is set by engine knock. The misfire limit is reached
when the GIMEP coefficient of variation (COV) is higher than 10%, and the knock limit is reached when
the Knock Intensity is over 3 MW/M
2
.
0.0
8
0.1
0.2
0.4
0.3
8
Map 1.9 knock
7-
I
Map 1.5 knock
0
M p 1.5 misfire
6-
6
Unstable
W
0.0
5-
0.6
A
Haltermann 437 Gasoline, Octane No. 96
T intake 8o C
PFI 70 CA BTDC intake
RPM 1200
Ic
0.5
5
Map 1.2 knock
Map 1.2 misfire
Map 1.9 misfire
4
4
0.0
0.1
0.2
0.3
OA
0.5
0.6
EGR rate
Figure 5-1 Map of the test engine working range
From Figure 5-1 we see that the knock limit curves for different intake pressures all go up with the
EGR rate. This is because with more inert gas, the ignition delay of the in-cylinder mixture gets extended
108
as indicated by the RCM test results; thus the mixture is less prone to knock and the knock-limited load
increases.. The longer ignition delay, however, retards the combustion phase and enhances the misfire
propensity of the engine; therefore, the low load limit becomes more restrictive.
With higher intake pressure, the misfire limit can be pushed even lower, as the in-cylinder
temperature and pressure conditions after compression are more favorable for auto-ignition, and more
available oxygen inside engine cylinder would also advance the combustion phase. The knock limit is
improved by intake boosting as well. As reasoned in Chapter 3, this is because higher intake pressure
gives higher in-cylinder pressure, so: 1) the magnitude of pressure oscillation will be smaller due to the
fact that it's harder for the pressure wave to be developed with the same pressure rise rate; 2) from the
definition of knock intensity we know that given all other conditions fixed, this directly lower the knock
intensity. Using KI instead of MPRR as knock metric and high load limit constraints made it possible to
capture these effects. Therefore higher intake pressure expands the operation range for CAI engines
with proper EGR coordination.
Notice the misfire limit of the tests with intake pressure 1.9 bar in Figure 4-1 looks higher than that
of the tests with intake pressure 1.5 bar. This is due to the limited number of data points on the map.
Similarly it can be seen that the test data point with intake pressure 1.5 bar and EGR rate 0.3 is below
the intake pressure 1.2 bar curve.
For tests with intake pressure 1.2 bar, the misfire limit curve and knock limit curve get crossed when
the EGR rate reaches 35%. This gives a singular point on the working map at GIMEP around 6.3 bar, and
the status of this point is very unstable since it tends to either knock or misfire from cycle to cycle. This
is because with lots of exhaust gas induced into the fresh charge, a bit less fuel will cause misfire due to
the much retarded combustion phase, and since lots of air is replaced by the inert gas for high EGR cases,
a bit more fuel in the cylinder will cause engine knock due to relatively high equivalence ratio. For tests
with higher intake pressures, the knock limit curve hits the high EGR limit when the equivalence ratio
approaches stoichiometric equivalence ratio, which is an artificial set up to restrain the experiment
within the range of lean combustion for satisfactory NOx emission performance. The misfire limit curves
stop at the same EGR rate as the corresponding knock limit curves. Easy to imagine if the tests are
conducted beyond the current EGR rate limits, crossed knock limit and misfire limit curves can be seen
in these tests as well.
109
From above we can see that a maximum working range can be achieved with intake pressure of 1.9
bar, with GIMEP from 3.5 bar to 8.9 bar and EGR rate of 0 to 54% respectively. To have an optimal base
case for the largest possible variation space for other parameters in the stratification effect study, a
nominal GIMEP of 6 bar under the condition of intake pressure 1.9 bar is picked for all the future tests.
This corresponds to a total amount of fuel per cycle at around 16 mg.
5.1.2
Single Pulse Direct Injection
The simplest method for fuel stratification creation and adjustment is by varying the timing of a
single pulse direct injection. In this section, all the fuel is delivered via a single pulse through DI portal.
The start of injection (SOI) timing is varied from the beginning of compression stroke (0 "CA ABDC
intake) to around TDC of compression (180 0CA ABDC intake) until misfire happens due to too late fuel
injection. This is determined after comparing the tests results to those with DI timing during the intake
stroke, which showed no obvious variation but moderate decreasing of knock tendency possibly from
cooling effect of DI fuel. To understand the impact of EGR dilution, different levels of EGR (0, 20%, 30%,
and 40%) are applied across the DI SOI sweep. The temperature of the EGR gas is always kept the same
as intake gas before mixing. The injection pulse width controlled from the testing system master
computer is kept constant at 1000 microseconds to assure a consistent total amount of fuel inside the
engine cylinder for each cycle. Engine performance was measured at 1200 rpm and 1.9 bar MAP. The
intake temperature is fixed at 80 *C and the total amount of fuel is fixed at 16 mg as described above,
which corresponds to an excessive air coefficient (A) equal to 4 at 0% EGR and 2.3 at 40% EGR.
1.
Knock Performance
Figure 5-2 shows the variation of knock intensity of the engine under corresponding conditionsWith
early injections (SOI before 120 "CA), for all EGR rates the knock intensity is always lower than the
audible knock threshold. When the DI SOI is retarded to later than 120 "CA, KI starts to ramp up. When
it's later than 140 "CA, KI drastically shoots up and peaks at 160 "CA. On a linear scale graph, the
differentiation by EGR rate is not obvious for the early injection cases but can be observed in the
knocking cases.
110
0
I
20
40
60
80
100
120
140
MAP 1.9 bar, PEX 2.0 bar
RPM 1200
sI Single Pulse DI
EGR=0
-E
--- EGR=20
-aEGR=30
60 -.- EGR=40
160
180
2010
I 100
s0
a
60
C40
140
20
I
Knock limit value
0
20
40
80
60
100
120
140
160
20
180
200
Di Sol (*CA ABDC [intakel)
Figure 5-2 KI variation with SOI timing for high CR single pulse DI tests
Notice in the knock regime, the peak KI of the curve with EGR=0 is lower than curves with EGR>0,
which is counter-intuitive. This may be because the pressure transducer did not capture the severe
pressure oscillation in the no EGR case as the peak pressures is very closed to the device limit (170 bar).
To see the fuel stratification and EGR dilution impact clearer for the early injection cases, the knock
intensity variation is re-plot as in Figure 5-3 with a log-scale. Note that human ears perceive acoustic
energy flux also on a logarithmic scale.
0
40
80
60
100
120
140
160
180
200
MAP 1.9 bar, PEX 2.0 bar
RPM 1200
100
E
20
100
Single Pulse Dl
-nEGR=Oa'
-.EGR=20
10
10
-i-EGR=30
S --
1
EGR=4O
1
S
1,
0.1...
0.1
0.01
0.01
1E-..
0
20
..
40
60
..
80
100
..
120
140
..
160
180
.E-3
200
DI SOI (*CA ABDC [intake])
Figure 5-3
KI variation with SOI timing for high CR single pulse DI tests (log scale)
111
For early injection (SOI from 0 to 120 *CA), the fuel relatively well mixed at ignition and engine knock
performance is not sensitive to injection timing, although moderate increase of knock intensity can be
seen at relatively later injections. With SOI later than 120 "CA, knock intensity starts to increase in a
much faster pace with SO retard, and peaks at 160 *CA. This is an effect due to locally rich zones formed
right before ignition with late injections. When SOI approaches compression TDC and gets very late,
unstable combustion starts showing up and misfire eventually occurs. This could be due to the very late
combustion phase caused by late injection and limited available time for fuel to complete evaporation,
mixing, and chemical delay processes. As in-cylinder temperature decreases along with downward
movement of piston after TDC, with too late injection, the time needed for mixing and pre-ignition
reactions to occur could exceed what's available constraint by the decreasing in-cylinder temperature,
and eventually causes misfire. Before that happens, the combustion phase gets retarded with later SOI,
which will weaken the engine knock tendency. There is indeed differentiation among tests with different
EGR rates for DI SO earlier than 160 *CA, however much not for even later.
MPRR is the "driving force" of CAI engine knocks. It is a measure of the heat release rate and can be
used to link the engine combustion process with the knock performance. The MPRR of the tests for
single pulse injections are shown in Figure 5-4.
0
20
40
60
80
100
120
140
160
180
200
MAP 1.9 bar, PEX 2.0 bar
RPM 1200
30
Single Pulse DI
EGR=O
EGR=20
-n-m-a-.-
30
EGR=30
EGR=40
'\20
20
-10
10
0
20
40
60
80
100
120
140
160
180
200
DI Sol ('CA ABDC [intake])
Figure 5-4
MPRR variation with SOI timing for single pulse DI tests
112
For all the curves with a specific EGR rate, the dependence of MPRR on DI SOI is of the same trend:
while DI SI keeps retarding, MPRR first barely changes, and then starts to climb up when DI SOI is later
than 120 *CA. The increment of MPRR becomes drastic when it's later than 140 *CA, and the MPRR
peaks when DI S01 equals 160 *CA. SOI later than 160 *CA, MPRR drops down until unstable combustion
happens in the very late injection cases. The absolute value of MPRR is generally reduced by a higher
EGR rate, especially for SOI earlier than 160 *CA.
2.
Combustion Performance
To better understand the engine knock performance, analysis of in-cylinder combustion process
corresponding to the above tests are conducted. In Figure 5-5, the variation of CA10 and CA50 with DI
S01 are shown respectively.
200
195
10
0
20
40
60
80
100
120
140
160
180
MAP 1.9 bar, PEX 2.0 bar
RPM 1200
Single Pulse Dl
a
I
185
a
o.8
200
200
0
200.....
a19
19
20
40
60
80
100
........................
120
140
160
160
200
MAP 1.9 bar, PEX 2.0 bar
RPM 1200
Single Pulse DI
195
190
190
s
18
18a
185
18
160
ISO~/
6
13 176
190
.
(0i 17u-u
170
EGR=20
--
160
-:EGR=0
:
175.l
170
EGR=
170
16
165
166
G=0
165
160
-U- EGR=40
155
165
160..........................150
0
20
60
40
80
100
120
140
160
155
155
15ISO
1:0
200
200
0
Dl S0l ( CA ABDC [intake])
..........................
20
40
60
60
100
120
140
160
180
150
200
D S0l ( CA ABDC [intake])
Figure 5-5 CA10 and CA 50 variation with SOI for single pulse DI tests
It is clear that retarding DI SCl in the early injection tests doesn't impact the combustion phase
much, although moderate retarding can be observed, which might be from the fuel stratification. From
120 *CA on, the combustion phase is advanced with SI retarding until 160 "CA. With even later DI SI,
combustion phase gets retarded again until misfire happens for too late ignition timing. For tests with
different EGR rates, the combustion phase is varied in such a manner: 1) For SOI earlier than 160 *CA, 040% EGR gives around 5 *CA and 8 "CA span to CA10 and CA50, respectively. 2) For SI later than
160 *CA, not much difference is made by EGR rate. The difference of the EGR effect in different regimes
113
of DI SO is very likely caused by the different in-cylinder temperatures before ignition for these tests, as
from Chapter 4 we know that the effectiveness of EGR could vary with ignition temperature, and lower
ignition temperature normally induces weaker differentiation by different EGR rates.
The engine knock tendency closely correlates with combustion phase: with earlier combustion
phase, the heat release process is more severe, e.g. for the EGR= 20% tests, with CA10 progressing from
175 to 172.5 "CA (SOI 120 to 150 'CA), the MPRR drastically goes up from 15 to 35 MPa/ms, and the KI
increases from 0.1 to 30 MW/M 2 ; for SO= 90 *CA, with CA10 retarded from 176 to 181 *CA (EGR=0 to
40%), the MPRR is reduced from 16 to 5MPa/s, and KI drops from a bit higher than 0.1 to 0.009
MW/m 2.These could be resulted from the ignition temperature differences, as shown in Chapter 6.
Also notice the SI position for CA10 valley point shows up around 5 *CA earlier than that of the
MPRR peak value. This means combustion phase and MPRR does not necessarily vary in the same pace
with SOI, i.e. besides combustion phase difference induced ignition temperature differences, the fuel
stratification differences directly impacts the engine knock tendency as well.
Figure 5-6 shows the burn duration of above tests. Regarding SOI effect, it's not very sensitive in
early injection cases (SOI 0 to 120 "CA) although moderate reduction can be observed. For the regime
where KI exceed the knock threshold, the burn durations are shortened down to less than 2 *CA. This
corresponds to the drastically increased MPRR. The burn duration extends by higher EGR rate with a
3 "CA span for tests with SOI earlier than 160 *CA, and not much for too late SI.
0
20
40
60
s0
100
120
140
160
180
10
MAP 1.9 bar, PEX 2.0 bar
RPM 1200
Single Pulse DI
8
8
EGR=O
-m-an-s-
o
200
10
EGR=20
EGR=30
6 e-n- EGR=40e
6
00
2
2
0
20
40
60
80
100
120
140
160
180
200
DI SOI (0CA ABDC [intake])
Figure 5-6
Burn duration variation with SOI for single pulse DI tests
114
The NOx emission of available range is shown in Figure 5-7. High NOxemissions seen in the region
around DI SOI 160 *CA corresponds to the severe engine knock and therefore high temperature
combustion. However the particularly short burn durations in this regime seen from Figure 5-8 would
confirm the nature of the combustion is CAl rather than diffusion flame combustion. The decreasing NOx
emission in the very late injection tests confirms that these cases are of CAI combustion nature as well.
Fairly differentiated NOx emission level for curves with different EGR rates are evidence for the
temperature reducing effect of it.
120
1200
900
0
140
160
160
MAP 1.9 bar, PEX 2.0 bar
RPM 1200
Single Pulse Dl
-aEGR=0
-aEGR=20
-EGR=30
-aEGR=40
200
1200
900
00
x4
b600
300
120
a
140
160
a
300
180
200
DI SO (*CA ABDC [intake])
Figure 5-7
NOx emission variation with S0l for single pulse Dl tests
Pressure traces of single pulse DI tests are compared with each other and a pure PFI test under EGR
rate 40% in Figure 5-8 to visualize the effect of SOI retarding. For tests with SOl earlier than 120 "CA, no
significant changes can be observed. From 120 *CA on, the crank angle for fuel heat release induced
pressure rise is advanced with the retarded SOI, and the peak pressure keeps increasing as well. The
slope of the pressure rise due to combustion is also increasingly steeper. For tests with DI SOI later than
160 "CA, the crank angle for fuel heat release induced pressure rise shows up later and later with
retarded SOI, and the peak pressure also drops monotonically. The slope of the pressure rise due to
combustion becomes less steep as well. These phenomena are consistent with the earlier interpretation
about the effect of SO.
Comparing DI SOI 90 *CA and 120 *CA with the PFI test, it is obvious that early DI injections help on
retarding the combustion phase and slow down the heat release process (seeing from the slope of
pressure rises). This is saying that although the engine combustion and knock performance doesn't
115
improve as expected with more and more stratified fuel charge with Dl timing retarding, the directly
injected fuel could create certain stratification and contribute to the knock mitigation efforts. The
cooling effect of directly injected fuel might share the credits as well. Comparing SOI 165 *CA and
168 *CA tests with the PFI test, we can see although the combustion phasing is retarded by the late
injection timing possibly due to the hard limit on time needed for evaporation and mixing, the pressure
rise during combustion is steeper than the PFI curve, indicating a higher knock tendency. This means
heavy stratification could overpower the combustion phasing (ignition temperature) impact and cause
severe heat release. Very late injection tests have both later and more relaxed heat release processes
than PFI as seen from SOI 170 *CA and 171 CA.
160
-
150
PFI
+50190
140
SSoil
-SO
130
20
140
Solso
S0160
S*
120
'
N
-~
s110
100
90
I
~u'4
80
'U
70
170
1i0
175
185
190
19S
200
CA (ABDC Intake)
160
11- PFI
150
+Sol
140
165
* So 168
SOI170
130
* S01171
S01160
120
110
*'*
100
90
'K
70
170
175
180
185
190
CA (ABDC intake)
195
200
205
Figure 5-8 Pressure traces for CR19, EGR=40%, Single Pulse DI tests
116
In summary, we find that for CAI combustion with single pulse DI on high CR (19) engine:
1)
DI SOI timing variation creates different levels of fuel stratification, which impacts both the
combustion phasing and burn duration, and thereby the engine knock propensity.
-
In the early SOI regime (0 to 120 "CA), the combustion phasing, burn duration, and engine
knock tendency are not significantly varied by SOI retarding due to the fast mixing process.
-
In a later SOI regime (120 "CA to 160 0CA), heat release process get advanced and drastically
more severe due to the locally increasingly richer mixture zones.
-
In the very late SOI regime (later than 160 0CA), the combustion phase gets retarded and
knock tendency reduces with SOI retarding, showing a strong ignition temperature effect on
the heat release rate.
The test results show a strong correlation between combustion phase and engine knock
tendency.
2)
EGR has significant effect on in-cylinder temperature and retards the combustion phase for tests
with SOI earlier than 160 "CA, and it also reduces the heat release rate as seen from burn
duration comparisons. Both effects help mitigate CAI engine knock. The effect of EGR is not the
same in all the tests. As shown in Chapter 4, the inert gas dilution is indeed more effective in
extending the ignition delay in higher temperature regimes.
3)
Comparing with PFI test, early DI injection tests show a better knock performance, which is due
to cooling effect. For very late SOI tests, the variation of combustion phase and burn duration
shows different trend than the early SOI tests.
5.1.3
Partial Fuel Stratification
So far, we have found that fuel stratification created by single pulse DI does not benefit the knock
mitigation as expected (later SOl gives reduced knock tendency). The stratified charge either doesn't
have strong enough effect to enhance the sequential combustion, or worsens the engine knock
tendency by having rich zones formed for advanced ignition. This means that single injection tests may
not have created the desired charge stratification.
To explore the opportunity of fuel stratification strategy comprehensively, "Partial Fuel Stratification
(PFS)" [41] is adopted and tested by delivering the fuel via both PFI and DI portal within one engine cycle
with specified split ratio between the two. In this set of tests, the total amount of fuel per cycle is kept
117
the same as the previous section, while 60% of it is prepared and delivered into the cylinder through PFI
portal, and the other 40% is through DI portal. By doing this, the extremity of fuel stratification can be
reduced without reducing the nominal power output. All the operation conditions are also kept the
same as before. The timing of the PFI injection is fixed at 70 *CA BTDC intake, and SOI of the DI is swept
from 90 *CA BTDC compression to around 15 *CA ATDC compression. For the purpose of consistency,
the crank angle positions are refereed in ABDC intake as before.
1.
Knock Performance
Figure 5-9 shows the KI variation with DI SO. The variation of KI of PFS tests shows a very similar
trend as single pulse DI tests, except:
1)
The absolute value of KI in the knock regime is much smaller than (~50%) in single pulse DI
tests with same amount of EGR dilution.
2)
The regime of very late DI 501 is much extended for stable combustion with DI SOl as late as
15 "CA ATDC compression.
3)
EGR dilution differentiated the knock performance properly in the heavily knocking region.
90
100
110
123
130
140
150
160
170
180
19
M AP=1.9 bar, P ,=2.0 bar
RPM 1200
PFI 60% (70 *CA BTDC[intake]);
30 D140%
-UEGR=0
-EGR=20%
- EGR=30%
20 -U-EGR=40%
39
29
Knock limit
value
10
90
100
110
16
123
130
140
150
160
170
180
19
DI timing ('CA ABDCpntake])
Figure 5-9 KI variation with DI SOI timing for PFS tests
118
Similarly Figure 5-10 is re-plotted with log scale for the data in Figure 5-9 to show the details in early
and very late DI SO regimes. For each curve with specified EGR, the DI SOI shows little impact on KI for
the early DI SOI tests. With DI SOI later than 120 *CA, KI starts going up and becomes more drastic when
it's later than 140 *CA. KI peaks at DI SOl of 160 *CA and drops after that, which is very similar to single
DI pulse tests. However, with even retarded DI SOI later than 178 "CA, KI ramped up for a short period
and then dropped down again to a much lower level until as low as comparable to early DI SOI tests.
EGR shows obvious impact on reducing the knock tendency in early DI SOI regime, but not much for very
late ones.
90
100
110
120
150
140
130
160
170
180
190
0
MAP=1.9 bar, PEX=2. bar
RPM 1200
10 PFI 60% (70 'CA BTDC[intake]);
04
10
DI 40%
-EGR=O
-UEGR=20%
-a- EGR=30%
1
EGR=40%
-U0.1
0.1
0
0.01
------
90
100
0.01
--
110
120
130
140
150
160
170
180
190
0
DI timing ( CA ABDC[intake])
Figure 5-10 KI variation with DI SOI timing for PFS tests (log scale)
Figure 5-11 shows the MPRR variation with DI SOI for tests with PFI and DI fuel. the curves show
similar trend as the single pulse Dl. Not much difference is made by SI variation in the early injection
cases and heavy knock occurs for DI SOl between 155 *CA and 165 *CA. However, the value of MPRR is
much smaller in this region, e.g. the peak MPRR of PFS tests is around 25 MPa/ms, while that of single
SOI
pulse DI tests is around 42 MPa/ms, which is more than a 40% drop. The knock region in terms of
tests.
range is smaller as well, which is roughly 155~165 *CA compared to 140~170 "CA for single pulse DI
This confirms that the high MPRR in this regime is caused by the much higher effective equivalence ratio
is
right before ignition due to late injection. Also, for tests with SI later than 170 *CA, the MPRR
119
reduced to a much lower level than the early injection cases, and the combustion is stable even when
0
the DI SOI is retarded to as late as 195 "CA. The tests are stopped at SOI 195 CA because too much
energy is lost due to very late combustion phase and the increasing unburned HC, which will be shown
shortly.
Regarding EGR effect, a 6 MPa/ms span can be found created by EGR rate from 0 to 40% in the early
injection tests (SOI earlier than 120 *CA), while not much impact is shown with DI SOI later than 160 *CA.
0
For tests with SO between 120 and 160 CA, the EGR effect gradually diminishes with the retarded SO.
27MAP=1.9
24
18
E
a 15
0.
120
110
100
90
27
130
140
150
160
170
180
2 0
bar, PEX= . bar
190
*
RPM 1200
24
PFI 60% (70 *CA BTDC[intake]);
21
18
EGR=O
EGR=20%
EGR=30%
EGR=40%
-U-U-
w15
12
12
9
27
2
9
--------
6
6
U5
3
90
100
110
120
130
140
150
160
170
180
3
190
DI timing ("CA ABDC[intake])
Figure 5-11 MPRR variation with DI SOI timing for PFS tests
2.
Combustion Performance
The variation of CA10 and CA50 with DI SOI are shown in Figure 5-12. For early injections (SO before
120 *CA), CA10 doesn't vary with injection timing; for later injections, the combustion phasing gets
advanced until SOI reaches 150~160 'CA (higher EGR rate retards the valley position); and for very late
DI injection (SOI after 160 *CA), the ignition timing retards with 501 and CA10 could even be earlier than
DI SOl. This means fuel delivered via PFI has ignited before DI fuel is delivered. This could explain the
stable combustion for very late DI SOI tests, as in-cylinder temperature is pushed up by the heat release
of PFI fuel first in these tests. It's also seen that EGR dilution of 0 to 40% creates a 3 "CA span for CA10 in
almost the entire SOI sweep range. This is different from single DI tests, meaning the EGR effect is
indeed more effective with higher in-cylinder temperature, as illustrated in chapter 4. For CA50, the
120
general trend with SOt variation is very similar, yet the EGR effect is shown in tests with SOI earlier
than
160 *CA only, with around 5 *CA span for 0-40% EGR in the early injection tests and smaller span
in the
relatively later ones. Obviously, the engine knock tendency largely correlates with combustion phasing.
Notice the correspondence between KI and CA50 for SOI around 185 "CA.
Comparing with single pulse DI tests, CA10 shows very similar values for tests with SOI before CA10,
but smaller values for tests with SOI after CA10, meaning the CA10 is governed by PFI fuel for these tests.
Similarly, CA50 of PFS tests have very alike values for tests with SOI before CA10, but much larger values
for tests with SOI after CA10, meaning the total combustion time period is much extended.
90
100
110
120
130
140
150
160
170
180
190
90
182
182
180 .
o
180
100
110
120
130
140
150
160
170
180
190
204
204
198
19
0
a
178
E 192
178
_________
-
-u-EGR=O
- EGR=20%
9
192
EGR=30%
mEGR=4O%
176
176
014 <
186
'20
-- EGR=O
- -EGR=20%
__
&n~180
EGR=30%___U18
_
_
---
~
-
174
174
180
I
CA101
172.
.
90
.
100
.
.
.
110
120
.
130
.
.
.
140
150
.
160
A17
. 172a
.
.
.
170
180
190
90
DI timing (*CA ABDC[intake])
100
110
120
130
140
150
160
170
180
190
DI timing (oCA ABDC[intake])
Figure 5-12 CA10 and CASO for CR19, EGR=40%, PFI and Dl dual injection tests
The burn durations of the above tests are shown in Figure 5-13. For tests with SOI earlier than
140 *CA, no material difference is made by retarding DI SOl. However injection around 160 *CA gives
much shortened burn duration, and very late DI SOI (after 170 0CA) gives much longer burn duration up
to values comparable with conventional engines. For the tests with DI SOl from 140 "CA to 170 *CA, the
high MPRR and KI values are found, but the very short burn durations makes it unlikely to be flame
propagation but CAI combustion.
Figure 5-14 shows the NOx emissions. For tests with DI SOI before 140 *CA, the emission is negligibly
low. With DI SOI passes 140 *CA, it ramps up and goes up to as high as 1000 ppm at around 165 *CA.
Interestingly, there is a 5 *CA delay of the NOx peak relative to MPRR peak. With DI SOI retards to later
than 165 "CA, the emission starts decreasing and eventually reach a level of 0~200 ppm depending on
121
different EGR rates. The relatively high NOx for the very late SOl tests indicates the existence of locally
rich zones before ignition. Naturally, no EGR effect shows up in the early SOI tests as no significant NOx
emission is produced in all of them, but in tests with observable NOx emission.
100
90
30
--
25 P
-U-3-n-
110
120
140
130
150
180
170
160
190
30
EGR=0
EGR=20%
EGR=30%
EGR=40%
Region of high MPRR
has short burn
duration; unlikely to be
flame propagation
20
7
25
20
0
E
15
.)'
= 15
M'/
I
10
10
5
5
0
0
100
90
110
120
140
130
150
170
160
180
190
DI timing ("CA ABDC[intake])
Figure 5-13 Burn duration for CR19 PFS tests
100
90
110
120
130
140
150
160
170
180
190
1000 * MAP=19 bar
. P EX20 bar
800
-i
0
z
RPM 1200
DI 40%
800
0 PFI 60%;
600
600
-N-
EGR=O
-U-
EGR=40%
400
,
400
M
M.
200 L
200
0
0
90
100
110
120
130
140
150
160
170
180
190
0
DI timing ( CA ABDC[intake])
Figure 5-14 NOx emission variation with DI SOI timing for CR19 PFS tests
122
Figure 5-15 shows the GIMEP variation with DI SOl. It is clear that the GIMEP is well kept constant at
around 6 bar. For tests with too early combustion caused by locally rich zones, and too late combustion
for limited physical preparation time before ignition, the GIMEP is lower, showing worsened combustion
efficiency.
90
8
100 110 120 130 140 150 160 170 180 190
8
6
6
4
4
Combustion
too advanced
Combustion
too retarded
2
2
0
90
100 110 120 130 140 150 160 170 180 190
0
DI timing (*CA ABDC[intakel)
Figure 5-15 GIMEP variation with DI SOI timing for CR19 PFS tests
Pressure traces of above tests with EGR rate of 40% are presented together with a pressure trace of
a PFI test under the same operation conditions in Figure 5-16, Figure 5-17, and Figure 5-18.
190
180
170
160
160
PFH
(U
.0
0
U)
U)
0
0.
140 -
140
120 -
120
100
100-
80 -
60
SOI 160 CA
SOI 155 CA
SOI 150 CA
SO 1140 CA
SOI 90 CA
'
TI170
80
60
e
40
-
40
190
180
Crank Angle (*CA ABDC)
Figure 5-16 Pressure traces for CR19, EGR=40%, PFI and DI dual injection tests I
123
170
175
160
180
185
,
190
195
160
,
PF
140-
140
120-
120
2100U)
100 -
U)
80-
-5.1170CA
.
'
-80
SOI 173 CA
SOI 175 CA
SOI 178 CA
SOI 180 CA
60-
40
170
100
SOI 160 CA10
SOI 165 CA
175
I
II
180
185
60
40
195
190
Crank Angle ( *CA ABDC)
Figure 5-17 Pressure traces for CR19, EGR=40%, PFI and DI dual injection tests II
170
180
190
200
160
210
.
160
PFI
140-
SOI 180 CA
SOI 183 CA
SO[ 185 CA
140
120-
SOI 188 CA
SOI 190 CA
SOI 192 CA
120
100-
-100
80-
80
60-
60
U)
U)
40
170
180
190
200
40
210
Crank Angle (0 CA ABDC)
Figure 5-18 Pressure traces for CR19, EGR=40%, PFI and DI dual injection tests Ill
From Figure 5-16 we can see that PFS tests with DI SI before 140 "CA give later ignition timing than
pure PFI test, which indicates improvement of knock tendency for PFS strategy. The effect could be from
fuel stratification and/or fuel evaporation. For DI S01 after 140 "CA, ignition gets advanced and the heat
124
release dominated pressure rise slope becomes steeper. This is similar to the general trend of pressure
trace variation with DI SOI retard in the single pulse Dl tests, except the SOI range of tests with better
performance than pure PFI test is larger.
Figure 5-17 shows the pressure traces of tests with DI SOI from 160 "CA to 180 *CA. The curves
follow the same trend as in Figure 5-10 for single pulse DI tests, yet in a much moderate manner. This is
because the amount of fuel injected through DI portal is 40% as before, and the stratification extent is
much reduced for all the cases here. When DI SOI is retarded to 180 "CA, two bumps on the pressure
trace can be seen. This indicates two independent heat release processes inside engine cylinder.
In Figure 5-18, with pressure traces for DI SOI later than 180 *CA, the separation of the two heat
release processes are more and more evident with DI SO being constantly retarded. This is saying that
fuel delivered into engine cylinder by PFI is ignited first, with the DI fuel gets ignited in a later moment,
and therefore the whole heat release process is extended by this separation of the combustion events
of PFI fuel and DI fuel. Here we can see that an extended heat releasing process can be achieved by
delivering the fuel into engine cylinders in relatively discrete time periods and having the combustion of
these fuel happen in a sequential manner. Notice effective power output might be reduced by too late
combustion, so specific goals and constraints are needed for the optimization.
To see and compare the heat release process of tests with SOI in different regimes more clearly,
Figure 5-19 shows the burned fuel mass fraction against crank angel for tests with 40% EGR. The result is
obtained via a Rassweiler and Withrow approach.
1.2
C
0 0.8
E0.6
0.4
2 0.4
-- W0 90 *CA
S-W 160 *CA
-SO1
0.2
178 CA
-501185 *CA
0
165
170
175
185
180
190
195
200
Crank Angel (ABOC intake)
Figure 5-19 Heat release process comparison for CR19, EGR=40%, PFI and D1 dual injection tests
125
Clearly the heat release process is advanced and shortened significantly from SOI at 90 'CA to
160 "CA, showing a strong rich zone effect. The severe oscillation of in-cylinder pressure can even be
inferred from the heat release curve of SOI at 160 *CA. Later than 160 *CA, the separation of the
combustion events of fuel delivered by PFI and DI starts showing up and becomes more and more
obvious with retarded SOL. It's also observable that the start of combustion of SOI 90 "CA is later than
the other three cases, this supports that smaller effective equivalence ratio might be achieved by early
DI, or the fuel cooling effect might impact.
In summary, the results from PFS tests on a high CR CAI engine show that:
1)
SOI Effects:
-
With early DI injection, the combustion is not sensitive to DI injection timing because the
fuel is relatively well mixed in the charge..
-
With later DI injection, the DI fuel is increasingly more stratified. As the timing is retarded,
the stratification of the DI fuel interacts with the PFI fuel: KI, MPRR, and NOx emissions first
increase because of ignition of locally higher fuel concentration regions are formed by the DI
fuel, and then decrease because combustion of the DI fuel and the PFI fuel separate into
two stages.
-
With very late DI injection, the DI fuel combusts in a HCCI mode separately from the PFI fuel,
producing low KI, MPRR, and NOx emissions. The separation of combustion could reduce the
engine knock tendency, yet the side effect of reduced effective power output along with it
should be noticed as well.
2)
EGR impacts CA50 in a similar manner as in single pulse DI tests that effects could only be seen
in tests with DI SOI before 160 *CA. However it differentiates CA10 curves in a more broad range
including the late injection tests (SOI after 160 "CA).
3)
Comparing with single pulse DI tests, the stratification extent is reduced, as can be seen from
the reduced values of MPRR, KI, and NOx in the heavy knock regime, and the reduced range of
heavy knock in terms of SOI.
4)
Comparing with PFI test, the PFS test with early DI SOI still show better performance, possibly
due to the more or less fuels stratification enhanced sequential ignition, or fuel cooling effect
from DI fuel evaporation.
5.2
Responses to Stratification at Compression Ratio = 15
126
From section 5.1 we see that fuel stratification can be created by directly injecting fuel into the
engine cylinder, and its extent can be varied via both DI SOI and PFI/DI ratio variation. Locally rich zones
are formed in tests with single pulse DI SOI around 20 'CA BTDC and lead to too early ignitions which
causes severe engine knock. With early DI injections, the engine combustion and knock tendency are not
sensitive to DI SOI variation in either single pulse Di tests or dual injection PFS tests, although
improvement compared to pure PFI test is observed. Substantail positive effects of fuel stratification on
CAl engine knock mitigation are not shown for the high CR engine tests. This could be resulted from the
insensitivity of fuel mixture ignition to the fuel stratification under the high temperature (~900 K) inside
engine cylinders after compression.
To verify this, tests on a lower compression ratio engine were conducted. The same facility was used
to keep all the parameters fixed except the compression ratio. A lower compression ratio of 15 is
obtained on the same engine by inserting an extra gasket between cylinder head and engine block.
5.2.1
Working Map of the CR15 Engine
From preliminary tests it is found that with the new compression ratio, combustion wouldn't occur
with intake pressures lower than 1.8 bar for all EGR rates at intake temperature 80 *C. Operating at
higher intake pressures than 1.8 bar, however, would shorten operation time of the super charger (the
limitation is set by the temperature of the motor running the compressor. Therefore intake pressure
was kept constant at 1.8 bar for the experiments with the lower compression ratio engine.
The typical working range of low CR engine is shown as in Figure 5-24. The engine speed and intake
temperature are fixed at 1200rpm and 80 *C respectively. The exhaust pressure is fixed at 1.9 bar. The
nominal EGR rate is swept from 0 to 35%. Same as the previous experiment, for each EGR rate condition,
the amount of fuel injected is varied to probe the load limits. The fuel is delivered via early DI injection
(30 "CA ABDC intake) to achieve thorough mixing of fuel and air, while the fuel cooling effect is kept for
better performance. As previously, the low load limit is set as GIMEP COV at 10%, and the high load limit
set by engine knock is hit when the knock intensity reaches 3 MW/m
2
.
From Figure 5-20 we see that with the portion of exhaust gas introduced into the intake varies from
0 to 35%, the knock limited GIMEP of the engine increases from 11.3 bar to 11.9 bar, and the misfire
limited GIMEP also raises from 7.8 bar to 10.2 bar. The A value could be as low as 1.5 for the highly
diluted case at knock limit, and as high as 3.0 for tests without EGR at misfire limit. If EGR is further
127
increased, combustion becomes very unstable, similarly as in the high CR map. Comparing the working
range of low CR engine with the high CR engine, we found that:
1)
The ignition conditions is harder to achieve for low CR as in-cylinder temperature after
compression is much lower than the high CR setup;
2)
The proper working range covers higher power output as both misfire limit and knock limit
shift up on the GIMEP-EGR map due to the lower in-cylinder temperature after compression;
3)
The behavior of knock limit and misfire limit of the low CR engine follows the same pattern as
the high CR engine.
Therefore a nominal GIMEP of 10 bar is set as the base case for a maximized parameter variation
space. This requires a total amount of fuel per cycle of 23 mg.
0.05
0.00
0.10
0.15
0.25
0.30
12-
0.35
-A
0.40
12
A
A
U
10
10U
0-
8-
8
7
U
-I.
i
0.00
i
0.05
i
i
0.10
0.15
i
0.20
0.25
0.30
0.35
1
0.40
EGR Rate
Figure 5-20 Working range of the low CR engine (RPM1200, intake T=80 "C, P=1.8 bar)
5.2.2
Low CR PFS Base Case
To keep the knock intensity within certain level for heavily knocking tests, fuel is delivered via both
PFI and DI portal within one engine cycle as before, and the fuel split ratio of PFI/DI=60/40 is adopted
for the base case test. Fixing the engine speed at 1200 rpm, intake temperature at 80 *C, intake pressure
at 1.8 bar, exhaust pressure at 1.9 bar, and the total amount of fuel per cycle at 23 mg, the DI SOI was
swept from early compression stroke (30 *CA ABDC intake) to 20 *CA ATDC compression, with EGR rate
varied in the entire SOI sweep range.
128
1.
Knock Performance
Knock intensity of the base case tests is shown in Figure 5-21 in log scale. The difference between
the shape of the curves here comparing with high CR tests is that for all the curves (except EGR = 20%)
with different EGR rates, there is a valley shape in the section with SI between 30 *CA and 120 *CA,
although the extent is different depending on EGR rates. This is saying that the engine knock tendency is
reduced by directly injecting 40% of the fuel in the middle of compression stroke (~105 *CA). The knock
2
regime in terms of S01 is around 150 to 170 "CA. For tests with no EGR, the KI drops from 0.6 MW/M to
0.05 MW/M
2
with DI SI retarded from 30 to 105 *CA, which is an order of 10 fold decrease. For tests
with 15% EGR, the KI does not vary much. The EGR effect is more obvious in early DI SC0 tests. Notice for
the tests with EGR=20%, there is no data available from S01 60 to 120 *CA due to misfire. This is due to
the extremely retarded combustion phasing, which will be shown shortly. Also notice the peak KI in the
2
map could be as high as 100 MW/M
for DI S0l at 160 *CA.
40
20
80
60
100
140
120
160
180
200
100
100
Knock Limit Value
10
10
---
-
j7
/-.-EGR=O%*
m-EGR=10%
U
0
*
0.1 SE
0.01
u
-
W
I-
-
EGR=1O%
EGR=15%
--
EGR=20%
-.
0.1
-0.01
20
40
60
160
140
120
100
80
0
DI timing ( CA ABDC Intake)
180
200
Figure 5-21 KI of the base case on low CR engine (RPM1200, intake T=80 C, P=1.8 bar)
Figure 5-22 shows the MPRR variation corresponding to the above tests. The MPRR curves also have
similar pattern as in high CR tests with late DI SI (120 *CA~180 "CA), but very different responses for
early DI SCI. Differently from KI curves, the effect of SI in the middle of compression stoke is more
significant for higher EGR rates. While nearly premixed charge induced by SI at 30 'CA with no EGR
129
gives an MPRR of 19.9 MPa/ms, 15% EGR drives it down to 9.9 MPa/ms, which is more than 50%
reduction. By retarding the DI SOI to 90 *CA, MPRR could be further reduced to around 2.0 MPa/ms,
which is a further 80% decrease, and also a nearly 90 % improvement of the engine knock tendency (in
MPRR) comparing to the test with DI SOI at 30 0CA and no EGR. When the EGR rate reaches 20%, misfire
happened in tests with DI SOI around the middle of compression stroke (60 "CA ~140 'CA ABDC intake).
20
60
40
80
100
120
140
160
180
200
40-
4fl
I4
EGR=O%
-.-
.
.*
30-
EGR=10%
EGR=15%
EGR=20%
30
20-
20
1
-
10-
20
10
40
60
80
100
120
,
140
160
180
.
|
200
DI timing ("CA ABDC Intake)
Figure 5-22 MPRR of the base case on low CR engine (RPM1200, intake T=80 *C, P=1.8 bar)
2.
Combustion Performance
Figure 5-23 shows the CA10 and CA50. Easy to see that for all the constant EGR curves with retarded
DI SOI, both CA10 and CA50 first retards, then advances, and then retards again. We can also see that
higher EGR has more significant combustion phase retarding with DI SOI in the middle of compression
stroke. With no EGR and DI SOI at 30 "CA, CA10 is at 182 *CA ABDC. By 15% EGR dilution, it is delayed to
184.2 *CA, and by retarding the DI SOI to 105 "CA it is further retarded to 189.6 *CA, which is in total a
7.6 "CA difference, indicating a decrease of the ignition temperature in the order of 50K. For CA50, the
test with no EGR and DI SOI at 30 *CA gives 183 *CA, while 15% EGR retards it to 187 *CA, and DI SOI at
105 *CA further pushes it to 194.5 *CA.
130
20
192
60
,
.
40
,.
.
0
,
100
. .
120
140
160
. . . .
180
20020
192
205-.
40
60
80
100
190s
S
8
184.
Is2
1O
-1218
IN
*
m184
180-m- EGR=0%
-EGR=10%
U
-a-EGR=15%
178 -r-EGR=20%
176
,
2
,
40
,
60
80
160
140
160
180
200
9
8
160
1ee
17
,
176
,17,
100
120 140
160
DI timing (CA ABOC Intake)
180
200
200
206
-EGR=O%
- EGR=10%
-EGR=15%
188 - - EGR=20%
-200
18801
120
20
40
60
,
..
80
100
120
140
160
DI timing ('CA ABDC Intake)
180
200
Figure 5-23 CA10 and CA50 of CR15 base case tests
The above observation indicates that early Dl in the compression stroke could actually retards the
combustion phasing, opposite to what was reported by previous research work [26, 41]. This could
be a
combination effect of fuel stratification and the generally delayed ignition preparation
process
(evaporation, mixing, chemical delay). By generally delayed ignition preparation, we mean with later SOI,
the time available for this process becomes shorter, and delayed ignition might be obtained before
the
locally rich zones overpower this effect.
The data gap on the EGR=20% curve is due to unstable combustion induced by extremely retarded
ignition timing at the higher EGR rate. Also notice the EGR effect is mostly seen in tests with
DI SO
earlier than 160 *CA, and it's more obvious in the middle of the compression stroke.
The burn durations for above tests are shown in Figure 5-24. It can be seen that for tests with no
significant EGR dilution (0 and 10%), there is moderate extension of burn duration by having DI
SOI in
the middle of the compression stroke comparing with very early injections (less than 1 "CA). However
for EGR=15% the burn duration is extended for around 6 *CA by retarding the DI SOI from 30 *CA to
105 *CA. This shows the direct impact of fuel stratification on burn duration, besides the indirect impact
via combustion phasing. Since the major effect of EGR is reducing in-cylinder temperature, this tells us
that the effectiveness of fuel stratification can be impacted by EGR as the in-cylinder temperature might
shift to regimes with different sensitivity of ignition behavior to equivalence ratio.
131
20
40
60
80
100
120
140
20-
--
160
180
200
20
EGR=0%- EGR=100
- EGR=15
1 EGR=2
16 -
- 15
10
- 10
C
0
5
6U
0
20
-
40
N
50
50
120
100
0
20 0
180
160
140
0
DI timing ( CA ABDC Intake)
Figure 5-24 Burn Duration from CR15, PFI and DI dual injection tests
To see the phenomenon more vividly, the pressure traces of the DI SO[ sweep tests with EGR=15%
are shown in Figure 5-25, Figure 5-26, and Figure 5-27.
175
110-
100.0
180
185
195
190
15% EGR
200
205
120
30 CA
-I
SSO 45 CA
SCII 60 CA
01i 75 CA01 90 CA
01 105 CA
S01 retard
110
- 100
0)
U)
U)
0)
0.
90 -
90
80 -
80
70-
70
60
175
180
185
195
190
200
205
Crank Angle (0 CA ABDC)
Figure 5-25 Pressure traces from CR15, EGR=15%, PFI and Dl dual injection tests I
132
170
140-
175
180
185
190
195
200
205
210
140
SOI 160 CA
15% EG
130-
SOI 150 CA
01I135 CA
501 128 CA
01 120 CA
01105 CA
120
110-
130
120
110
100-
100
909)
.
90
80 -
80
70-
70
So retard
60
50
170
175
180
185
190
- 60
1
1
195
200
205
, 50
210
Crank Angle (0 CA ABDC)
Figure 5-26 Pressure traces from CR15, EGR=15%, PFI and DI dual injection tests 11
170
14
175
,
180
,
195
200
,
,
-O
120
11-Sol
S
190
15% EG
130
100-
185
205
210
,
,
140
01160CA
130
1170 CA]
01175 CA
111180 CA
120
0'185 CA
0188 CA -110
501190 CA
110I
- retard
100
90-
90
80-
80
70-
70
60-
60
50
,
170
,
175
180
,..
,..
185
190
Crank Angle (*
,
C
195
---. ....
.....
50
200
205
210
A ABDC)
Figure 5-27 Pressure traces from CR15, EGR=15%, PFI and DI dual injection tests IlIl
Heat release analysis for selected tests is conducted and shown in Figure 5-28. The
curves
corresponds to tests with EGR=15%.
133
1.2
1
0
L 0.8
U-
C
' 0.6
DI SOI 30 CA
-DI SOI 105 CA
20.4
-DI
Sal 160 *CA
DI SO 180 *CA
U6
0.2
-D
SOI 188 *CA
0
172
177
182
187
192
197
202
207
Crank Angle (deg ABDC intake)
Figure 5-28 Heat release process comparison for CR15, EGR=15%, base case tests
With DI SOI being retarded, the combustion phase first gets retarded, then advanced, and then
retarded again. Along with this, the peak pressure also decreases from 115 bar to 80 bar with DI SOI
retarded from 30 *CA to 105 *CA, then it increases to 135 bar when DI SOI reaches 160 "CA, and then
decreases again to 67 bar with DI SOI at 190 CA, which actually represents the pressure peak created
merely by the 60% fuel delivered via PFI portal. The pressure rises induced by fuel combustion have
slopes that flattens out in the early DI injection section, steepens up in the later DI injection regime, and
then flattens out again with very late Dl injection.
It is clear that with retarded DI SOI, the heat release process is much extended and relaxed from
30 *CA to 105 *CA. With even later DI timing until SOI 160 "CA, the heat release process is both advanced
and steepened. Later than that, the curves show more and more obvious two sub-heat release
processes within one engine cycle, which makes the entire process a lot less intense. It can be seen that
the effect of fuel stratification formed by having DI SOI in the middle of compression stroke gives very
comparable heat release process with the very late injection induced dual combustion process. However,
the power output of these tests are kept well around the nominal 10 bar GIMEP, while the very late DI
SOI tests show a reduced GIMEP, as can be seen form Figure 5-29.
134
14
-EGR=0
-- EGR = 10%
*EGR
12
= 15%
-0-EGR = 20%
'U
8
Power
output well kept
Reduced power output
6
20
40
60
80
100 120 140 160
DI SOl ( deg CA ABDC [intake])
180
200
Figure 5-29 GIMEP comparison for CR15, EGR=15%, base case tests
In summary, the results from low CR PFS base case tests show that:
1)
SOI Effects: on low CR engine, by having DI SOI in the middle of compression stroke, the engine
knock tendency can be greatly reduced without power output loss, showing the effect of fuel
stratification and possible effect of the delayed ignition preparation time. The different pace of
KI and MPRR variation, especially for the high EGR (15%) tests, indicating that sequential
combustion is dominating and no pressure wave is developed inside engine cylinder.
2)
There is discrepancy between the combustion phasing variation and the burn duration variation
for the 15% EGR test with DI SOI in the middle compression stroke, indicating a direct effect
from fuel stratification to the burn duration besides the indirect one from combustion phasing.
3)
EGR shows significant impact on the effectiveness of the DI SOI variation on engine performance.
Higher EGR makes the mid-compression stroke DI SOI effect more effective, yet too high EGR
rate could cause unstable combustion or even misfire.
5.2.3
Fuel Split Ratio Impact
Experimental tests were conducted with PFI/DI fuel ratio varied on the low CR engine to explore the
optimization of fuel stratification creation and control for CAI knock mitigation. PFI/DI ratios of 40/60
and 80/20 are tested respectively and compared with the base case (PFI/DI=60/40). The total amount of
fuel per cycle and all the other operation conditions are kept the same for these tests as in the base case.
135
The GIMEP of PFI/DI=40/60 and 80/20 are shown in Figure 5-30 first to give a big picture of the
operation of these two sets experiment. From the figures, different fuel split ratios give different
tolerances to EGR for tests with DI SOI in the middle of compression stroke. With higher level of
stratification by splitting the fuel in 40/60, 10% EGR works for the entire DI SOI map, while 15% EGR
causes misfire for DI SO between 60 "CA and 120 "CA. With even higher EGR at 30%, combustion could
only be facilitated for DI SO1 around 160 *CA. However with a smaller fuel split ratio 80/20, the EGR rate
could go as high as 20% for full coverage on the map, although with even higher EGR at 25%, misfire also
happens for DI SOI between 60 "CA and 160 *CA. This serves as a cross evidence with 501 variation effect
to show that fuel stratification could indeed weaken the auto-ignition and abet CAI knock tendency.
20
40
60
80
100
120
140
160
20
180
40
60
80
100
120
140
160
180
200
12
16
"
14
*
EGR=O
EGR=10%
PFI/DI =40/60
EGRI15%
EGR=30%
U
4
14
EGR=O
*EGR=1O%
EGR=15%
PFI/DI =80/20
11
UEGR=20%
*EGR=25%
12-
10 -
10
9-
9
8
8-
8
6
7
12
Lu
.0
I0
LU
-
10 -
~
- 10
0.
LU
C,
8-
20
40
60
60
100
120
140
160
7
20
180
40
60
DI timing (*CA ABDC Intake)
80
100
120
140
150
180
2
DI timing (*CA ABDC Intake)
Figure 5-30 GIMEP for PFI/DI ratio of 40/60 and 80/20
1.
Knock Performance
Knock intensity are compared among different fuel split ratios as in Figure 5-31 for EGR=0 and 15%
respectively. First look at the EGR= 0 graph, we find with different fuel spilt ratio, the variation of KI with
DI SOI basically follows the same trend. However, three differences among them should be noticed:
1)
The KI value in the early SOI regime (-before 100 *CA) increases with the PFI/DI ratio,
meaning the more fuel is delivered by PFI, the higher knock tendency is. This is evidence that
early direct injection also reduces the CAI knock tendency.
136
2)
The mid-compression stroke SOI effect is more obvious for tests with lower PFI/DI value,
meaning stratification effect on CAl knock mitigation is enhanced with more fuel delivered by
DI. This is a cross evidence with 1) for the fuel stratification effects on CAI knock mitigation.
3)
The position of KI value drastically ramping up (in terms of SOI) advances with lower PFI/DI
value, meaning the stratification effect on forming locally rich zones are strengthened by
more DI fuel as well. Notice the absolute value of KI at the same S01 position for different
PFI/DI curves are also higher for tests with more DI fuel.
Compare the EGR=15% graph with the EGR=0 graph, we find very alike trend of the curves, while the
absolute value of KI is much lower due to the EGR effect. Also notice the variation of KI in the early
injection regime is not as significant as in EGR=0 tests. As mentioned before, this is because no pressure
wave is developed under such conditions and absolute value of KI should have no physical significance.
The PFI/DI=40/60 curve has data gap for SOI between 60 and 120 *CA due to misfires, which confirms
the fuel stratification effect on combustion phasing retarding even further. It also has data gap for SOI
later than 170 *CA due to misfire, showing consistent characteristic as the single pulse DI tests with CR19.
20
40
60
80
100
120
140
160
180
20
200
40
80
0
100
120
140
160
180
100-
PFI/DI=40/60
PFI/DI=60/40
* PFI/DI=80/20
EGR=O
-0
100
100
100
PFI/DI=40/60
PFI/DI=60/40
- PFI/DI=80/20
EGR=15%
U
E
10
200
10
0.1-
%
O'
u.1-u-
U
-1
.-
0.01
0.01
20
40
60
80
100
120
140
160
DI timing (*CA ABDC Intake)
180
m--
1
0.01
U
12 -3
200
20
40
60
80
100
120
140
160
DI timing (*CA ABDC Intake)
180
1E-3
200
Figure 5-31 KI for PFI/DI ratio study with EGR=O and 15%
Figure 5-32 shows the MPRR comparisons of above tests. Easy to see that MPRR in both graphs
follow the same trend as KI, except with higher EGR dilution, the mid-compression stroke SOI effect is
more material, indicating even enhanced sequential combustion in these tests.
137
CU
-U40-
w
-
UU
12
1U
160
1
100
200
PFI/DI=40/60
PFI/DI=60/40
PFI/DI=80/20
120
40
140
20
160
40
0-
i-
9
-
EGR=O
30
20
20
30-
20
0.
200
80
PFI/DI=40/60
*
30 -X
180
60
-
40
PFI/DI=60/40
PFI/DI=80/20
30
EGR=15%
20
--
0.U
u
-
10
-10
10
0
0
0
20
40
so
80
Di
100
120
140
180
180
200
10
_T
20
timing (*CA ABDC Intake)
40
60
80
DI
100
120
140
160
.
.........
.........
0
180
200
timing (*CA ABDC Intake)
Figure 5-32 MPRR for PFI/DI ratio study with EGR=O and 15%
Therefore, the fuel split ratio impacts the fuel stratification comprehensively: more fuel delivered
via DI would strengthen the mid-compression stroke DI SOI effect on knock mitigation, but will also
enhance the drastic pressure rise in late SOI tests, and increase the risk of misfire for both midcompression stroke DI SOI tests, and very late DI SOI tests respectively. This means neither too heavy
nor too light stratification is good for CAI engine knock mitigation, as too extreme stratification
encounters misfire and drastic pressure rises easily, and too light stratification doesn't have enough
sensitivity to effectively impact.
2. Combustion Performance
Combustion phase analysis is conducted for the tests above. Figure 5-33 shows the comparison of
CA10. It can be seen that CA10 correlates to the variation of MPRR in Figure 5-32 very well. Look at the
EGR=0 graph, we see that all three curves show very similar pattern, except:
1)
Lower PFI/DI tests have stronger effect on CA10 retarding, meaning the effective equivalence
ratio before ignition is lower for the same SOI.
2)
Lower PFI/DI curves has more advanced start SOl position of the ignition timing advancing due
to local rich zones formation, and more advanced SO position of the valley CA10, although the
absolute value of CA10 is not the minimum. This also shows enhanced stratification effect.
3)
Lower PFI/DI curves has less retarded CA10 in the late SOI DI regime (after 160 *CA). This is
naturally because the with more DI fuel, there's less PFI fuel, and for too late DI SO, CA10 is
138
governed by the PFI fuel auto-ignition preparation. Also the misfire position of SOI is earlier for
tests with more DI fuel, consistent with the single pulse DI tests on high CR engine set up.
Looking at the EGR=15% tests, we find that EGR generally retards the ignition timing, while it also
enhances the DI S01 effect in the mid-compression stroke. Easy to see, with no EGR, the difference
among curves with different PFI/DI ratios is in the magnitude of 1 *CA. However with 15% EGR, the peak
value of CA10 can be of a 3 "CA difference between PFI/DI 60/40 and 80/20, and the 40/60 case even
misfired for too retarded CA10.
20
40
60
80
100
120
140
160
180
20
200
40
60
80
100
190 18
-
U
184 -
U
4
4
U
0
4
U
140
180
166
200
190
I
PFI/ DI=40/60
PFI/ DI=60/40
PFI/ DI=80/20
-
180 -
*
*
,
40
60
-188
184-
182
- -188
184
182-
-C
*/
180
-"
"
U
, -,
20
-
U
EGR=0
178-
-
184
I-
182 -
176-
188
El
3
.16
a
120
*
186
.
(2
U
,I-
.
.
U
178
..,
.
80
100
120
140
160
DI timing (*CA ABDC Intake)
180
178 -
1SO
EGR=15%
178
170-
176
182
PFI/DI=40/60
PFI/DI=60/40
PFI/DI=80/20
200
20
40
80
DI
80
100
120
140
166
timing (*CA ABDC Intake)
176
1-
180
200
Figure 5-33 CA10 for PFI/DI ratio study with EGR=O and 15%
The trend of combustion phase is further reflected by CA50 in Figure 5-34.
20
40
60
80
100
120
140
160
180
200
40
20
80
60
100
120
140
160
180
200
190
200
200
m
190
195
a
185
185
*
U
-m*
-
O180
190
-
185
-
190
M
0
PFI/DI=40/60
PFI/DI=60/40
PFI/DI=80/20
U,
_-
a
*
-
0
0
180
EGR=O
20
40
60
*W\*
80
100
120
140
160
160
200
0
1
EGR=1 5%
180-
20
DI timing (*CA ABDC Intake)
185
PFI/DI=40/6 0
PFI/DI=60/4
PFI/DI=80/2
40
60
180
.
80
I..
100
I.
120
T
140
-.
160
DI timing (*CA ABDC Intake)
Figure 5-34 CA5O for PFI/DI ratio study with EGR=O and 15%
139
---
I.
180
....
1
200
With PFI/DI=80/20, the latest CA50 achieved in the tests with EGR=0 and 15% is 186 *CA and
191 *CA respectively, whereas with PFI/DI=60/40, the latest CA50 is 187 *CA and 195 *CA. For
PFI/DI=40/60, CA50 is retarded to as late as 187 *CA in the test without EGR and misfire happened due
to too late combustion phase in the EGR=15% test.
Burn durations are shown in Figure 5-35. For EGR=0, no significant difference is seen for different
fuel split ratios. The curves follow the general trend that with DI SOI retarding, the burn duration slightly
extends (30 to 105 *CA), then shortens (105 to 160 'CA), and then greatly extends (after 160 *CA). This
means fuel stratification under such conditions doesn't directly impact the burn durations. However
with EGR=15%, the burn duration is differentiated in the majority SOI regime by fuel split ratio. As large
as 4 "CA discrepancy is found between the 60/40 and 80/20 curves with DI SOI at 105 *CA. The
difference made by DI SOI is also obvious, indicating a direct impact of fuel stratification on burn
duration by sequential combustion enhancement besides indirectly via combustion phasing retarding.
20
60
40
80
100
120
140
160
180
20
200
U
- 15
100
120
140
160
180
12
E
<
- 9
09
6
- 1s
U
EG=5
EGR=15%
12
-
6
-6
6
6
m
--
/
0'
40
120
140 160
60
80
100
DI timing ('CA ABDC Intake)
180
3
6
0-
1.0
20
12
- 9
'
09
-6
3
200
is
PFI/DI=60/40
1 -
- PFI/DI=80/20
EGR=O
12
80
-U- PFI/DI=40/60
PFI/DI=40/60
PFI/DI=60/40
PFI/DI=80/20
-i-
15 -
60
40
18
118
20
200
40
100
120
140
160
80
DI timing 0*CA ABDC Intake)
60
180
0
200
Figure 5-35 Burn Duration for PFI/DI ratio study with EGR=O and 15%
Therefore fuel split ratio has significant impact on fuel stratification creation and control. More Dl
fuel enhances the fuel stratification effects in all perspective: knock mitigation for mid-compression
stroke DI SOI, and drastic pressure rise for late DI SOI, as well as misfire for very late DI SOl.
To effectively utilize fuel stratification:
1)
Proper EGR dilution is needed to push the in-cylinder temperature into a regime with high
fuel ignition sensitivity to fuel stratification;
140
2)
The fuel split ratio needs to be optimized to create sufficient stratification for the knock
mitigation operation in the mid-compression stroke, yet not too extreme to induce misfire.
3)
The impact of fuel stratification on misfire and drastic pressure rise in different DI SOI regimes
should be noticed and taken care of.
In summary, fuel stratification can be used to effectively extend the burn duration and reduce the
CAI engine knock tendency, but subtle combination with EGR dilution and proper setting of stratification
formation via both fuel split ratio and DI timing needs to be configured. In the compared tests,
PFI/DI=60/40 gives the best potential of CAI engine knock mitigation.
5.2.4
Intake Temperature Impact
Intake temperature has significant influence on CAI combustion. The variation of intake temperature
is also extensively used in the operation of CAI engines for purposes including ignition enhancement,
combustion phase controlling, knock mitigation, and so forth [34,37]. Understanding the impact of
intake temperature to fuel stratification and EGR effects is of significant importance to the effort of CAI
engine knock mitigation. A comparison test set was conducted with all the other operation conditions
kept the same except intake temperatures.
With a PFI and DI dual injection set up on the low CR engine, intake temperatures of 60 0C, 80 0C,
and 100
0
C are tested for DI SOI sweeping and EGR sweeping orthogonally. The lower bound
temperature is set with the reference of condensation point for water vapor contained in the EGR
diluent. The upper bound temperature is set upon considerations of both the peak pressure limitation
inside engine cylinder and the realistic equivalence ratio range. This is because higher intake
temperature causes less intake air mass because of the lower air density. The fuel split ratio of 40/60 is
adopted for this study to magnify the stratification effects (more DI fuel gives stronger stratification
effects), while the total fuel amount is kept at 23 mg/cycle as before. All the other operation conditions
are kept the same as the base case test.
Firstly, the GIMEP of tests with different intake temperatures are shown in Figure 5-36 with two
different EGR rate. It is clear that the GIMEP is well controlled at around 10 bar for the majority tests.
However we find that with PFI/DI=40/60, the PFI fuel couldn't be ignited independently before the DI
fuel delivered, making stable combustion impossible for very late DI SOI beyond 180 'CA. In both graphs,
the 100
0
C intake temperature test curve is stopped at 140
141
0
CA, due to the very intense engine
vibrations for SOI even later. Notice in the EGR=10% graph, only intake temperature 80 "C and 100 *C
curves are seen. This is because with 10% exhaust gas dilution, ignition couldn't be triggered for any DI
SOI with intake temperature as low as 60 *C.
100
80
60
40
20
16-
.
-
140
160
1
Tin=80 'C
Tin=100 *C
14
14 -
12
12
EGR=O
12-
10
10o
80
120
100
140
160
180
16
14
Tin=80 'C
Tin=100 *C
EGR=10%
-
-
- 12
6
60
80
100
120
140
160
*
10
8--.
6
6
6
40
U
10-
8
-
a.
20
60
OLUL
IL
LU
40
20
16-
10
16
Tin=60 *C
-U14
~120
20
180
40
60
80
100
120
140
160
180
Di timing (OCA ABDC intake)
DI timing (*CA ABDC Intake)
Figure 5-36 GIMEP for intake temperature impact study, EGR=0 and 10%
1.
Knock Performance
Figure 5-37 shows the comparison of KI for above tests. From the EGR=0 graph we see that, with
intake temperature increased from 80 *C to 100 "C, the variation of KI with DI SOl shows a very different
trend: no obvious effect on knock mitigation can be observed on the 100 *C curve with DI SOI in the
2
entire map. The value of KI for high intake temperature tests are much higher as well, e.g. 30 MW/m
for 100 0C vs. 0.07 MW/M 2 at DI SO= 105 *CA. This means higher intake temperature deteriorates the
knock mitigation efforts significantly. In the other hand, lower intake temperature at 60 "C improves the
knock performance of the engine moderately, yet noticing the already very low KI level for 80 *C tests.
The increment of KI for later DI SOI at around 130 *CA shows higher values and steeper changes for
60 0C tests than 80 *C tests, implicating that lower temperature enhances the sensitivity of the fuel
ignition to fuel stratification.
Looking at EGR=10% graph, we found the 60 *C test data is missing, as the too low intake
temperature made it impossible for stable combustion to occur with any SOI set up. Comparing the two
graphs, we find EGR effectively reduces the KI values for the 80 *C tests but not likely for the 100 "C tests,
e.g. at 90 "CA, 10% EGR makes the KI of test with intake temperature 80 *C drops from 0.07 to 0.01
142
MW/M
2
(85.7%) while that of 100 *C is 5.5 to 3.0 MW/m
2
(45%). This is saying that the intake
temperature impacts the effectiveness of EGR dilution as well: higher intake temperature reduces the
EGR effect on CAI engine knock mitigation.
20
40
60
80
100
120
140
180
160
*
100
20
40
60
80
100
120
140
1010000
160
10
10
180
100
'
10
1
1
0.1
0.1
0.1
*
-
0.01
20
60
80
--
100
/
-Tin=60'C
0.01U
0
Tin=80 *C
--m- Tin=100'C
EGR=O
120
140
160
.
0.01
180
Tin=80 *C
w Tin=100'C
EGR=10%
U
1IE-3
20
DI timing (*CA ABDC Intake)
4:0
.
60
80
,
100
,
120
,
140
,
160
0.01
IE-3
180
DI timing ('CA ABDC Intake)
Figure 5-37 KI for intake temperature impact study, EGR=0 and 10%
Figure 5-38 shows the comparison of MPRR for the above tests. Easy to see the MPRR follows the
same trend as KI. In the EGR=0 graph, while intake temperature 60 "C reduces the MPRR by around 5
MPa/ms at DI SO=90 *CA from the base intake temperature, the intake temperature 100 "C increases it
by almost 20 MPa/ms. The non-linear impact of intake temperature is also shown on MPRR: the change
of MPRR between 60 *C to 80 *C is much smaller than that between 80 *C and 100 *C. This is very
possibly because with lower intake temperature from 60 *C to 80 "C, the ignition delay of the fuel-air
mixture locates in the NTC regime, and higher intake temperature up to 100 0C pushes the mixture
ignition delay into a more temperature sensitive regime. This is cross evidence for the argument that
lower intake temperature enhances the fuel ignition sensitivity to fuel stratification, as normally low
temperature sensitivity corresponds to high concentration/equivalence ratio sensitivity.
Comparing the two graphs with different EGR rate, we see the weakened EGR effect by higher
intake temperature reflected by MPRR as well. With 10% EGR dilution, intake 80 "C curve gains more
than 5 MPa/ms by 10% EGR dilution in most of the early DI SOI section (30 "CA to 120 "CA), yet intake
600C curve could only give around 2~3 MPa/ms improvement. This is saying for the same fuel density,
the sensitivity of charge ignition behavior to EGR dilution varies with temperature. In the tested case,
the sensitivity is stronger in the low temperature regime.
143
80
60
40
20
100
120
140
160
50*-
Tin=60
4
40--
30
30-
-
30/-
Tin=80 *C
Tin=100OC
120
140
160
20
0
60
80
100
120
140
160
1
10
10
10-
40
180
80
30
220-
Ur
20
-
100
a
-1
0
60
EGR=10%
EGR=O
(z
80
40
sso
C
Tin=10 'C
40
20
2-
180
50o
020
0
180
40
60
80
100
120
140
160
0
180
DI timing (*CA ABDC Intake)
DI timing ('CA ABDC Intake)
Figure 5-38 MPRR for intake temperature impact study, EGR=0 and 10%
Therefore high intake temperature worsens the engine knock tendency by diminishing both fuel
stratification and EGR effect, and low intake temperature enhances the fuel stratification effects,
although in all aspects.
The effect of intake temperature could be feasibly interpreted. Basically by having a higher intake
temperature, two things happen to the in-cylinder charge: 1) higher in-cylinder temperature. 2) less air.
Generally, higher temperature makes auto-ignition easier (unless charge is in the NTC regime). Less air
gives higher equivalence ratio for the same amount of fuel in unit volume, which enhances the autoignition as well. The impact of intake temperature on the sensitivity of fuel ignition to fuel stratification
is pretty evident. The impact of intake temperature to the effectiveness of EGR dilution is very likely
driven be the EGR effect on charge temperature, and the different sensitivity of fuel ignition to
temperature in different ignition temperature regimes, e.g. low sensitivity in NTC region, and high
sensitivity in low temperature regimes.
2. Combustion Performance
Figure 5-39 shows the CA10 of above tests. Increasing intake temperature advances the combustion
phase. From the EGR=0 graph we can see that from 60 *C to 80 'C, CA10 is advanced by 2 *CA in average
in the early DI SCl section (30 'CA to 105 'C), and from 80 'C to 100 *C, it's advanced by 3 *CA -6 *CA.
The fuel stratification with retard of early injection (300 - 90' abdc intake) caused combustion phase
retarding showing up on the 60 'C and 80 'C curves but not in the 100 'C curve. With the higher intake
temperature at 100"C, the ignition behavior of the in-cylinder charge after compression becomes much
144
less sensitive to fuel stratification. This insensitivity is resulted from the movement of the pre-ignition
that
chemistry from the NTC region into the high temperature regime. The graph of EGR=10% shows
the
with EGR dilution, the stratification effect of lower intake temperature tests is magnified, while
CA 10 of intake
higher intake temperature tests still doesn't show any fuel stratification effect. The
to 90 *CA, whereas it's
temperature 80 "C is retarded for near 5 *CA by delaying the SI SOI from 30 *CA
not changed in the intake temperature 100 "C curve. Comparing the two graphs we see that higher
intake temperature diminishes the effectiveness of EGR on combustion phasing retarding. In summary,
magnitude of the
intake temperature impacts the combustion phasing both through the effect on the
Figure 5-40 shows that
ignition delay and through the region where the pre-ignition chemistry occurs.
CA50 follows the same trend as CA10.
80
60
40
20
186-
160
140
120
100
-
-
189
109
186
185-
183
183
am
-
177
Tin=60 C
-E
)
7
80
60
40
160
180
189
-186
183
-
183-
<
U-
1
17
Tin=100*C
120
100
160
140
20
180
EGR=10%
60
40
174
80
100
120
140
160
160
DI timing ("CA ABDC Intake)
DI timing (*CA ABDC Intake)
Figure 5-39 CA10 for intake temperature impact study, EGR=0 and 10%
20
60
40
80
100
140
120
160
192-
U-1
18
U _
Tin=60 *C
Tin=80*C
-
20
180
192
40
60
80
100
120
140
160
185-
1
Tin=80 'C
Tin=100*C
--
112
189
180
195
192
18
1-EGR=10%
EGR=0
*
*-186
186-
186
S186-
0
183
1831883
<183
4
O
40
7
7
a
EGR=O
20
140
120
*Tin=80"C
Tin=80 *C
Tin=100*C
a
100
<
-
177-
cc
S
80
60
40
20
180
1
189-
80
SO
DI
timing
120
140
(*CA ABDC
Intake)
100
10
180
-
4
4
DI timing (*CA ABDC
Intake)
Figure 5-40 CA50 for intake temperature impact study, EGR= and 10%
145
so
10
Figure 5-41 is burn durations of the above tests. From graph of EGR=0,the burn duration is sensitive
to SOI at 600 intake temperature, with the burn duration first increases with retard of SOI and then
decreases. The sensitivity to SOI decreases with increase of intake temperature. At Tin=100' C, the burn
duration only decreases with retard of SO. The behavior at EGR=10% is similar.
20
40
8
60
80
100
120
140
=
160
20
10
60
40
100
s0
120
Tin=60 *C
Tin=80 *C
Tin=100*C
EGR=0
-U-
10-
-a-
U-
a-
a
-
U-
Tin=80 *C
Tin=100"C
10
EGR=10%
4
*
-1 s012
8
U
.2
.2
160
140
8
6-
6
4-
4
2-
2
a
2/U
0
0*
20
40
60
80
100
120
140
DI timing (*CA ABDC intake)
,
0-r20
0
18 0
160
.
40
,
60
.
, . ,
80
100
,
,
120
. ,
. ,
140
160
-18 0
160
DI timing ("CA ABDC Intake)
Figure 5-41 Burn Duration for intake temperature impact study, EGR=0 and 10%
Figure 5-42 shows the comparison of the 2,
values between tests with intake temperature 80 "C
and 100 "C. It can be seen that with higher temperature, the charge becomes much richer due to less air
induced into the engine cylinder. The drop of X value of intake temperature 80 "C with delayed DI
timing is unexplained.
20
3.0 v
40
60
s0
100
120
140
160
I s0
3.0
,
2.7-
2.7
2.4-
2.4
2.1-
2.1
1.8
1.5
* Tin=80 "C
a Tin=100*C
EGR=10%
.
12.
20
40
s0
80
100
120
140
, ,
160
1.2
Iso0
DI timing (*CA ABDC intake)
Figure 5-42 Excessive air coefficient comparison for tests with different intake temperatures
146
3. CAI Combustion with Higher Intake Temperature
To see the impact of increased intake temperature with more details, a full map of KI versus EGR
sweep and DI SI sweep at intake temperature of 100 *C is shown below in Figure 5-43. With intake
temperature at 100 "C, majority curves doesn't show reduced KI in the entire map comparing with the
near-premixed test (DI SI at 30 'CA). With 25% EGR, KI was reduced from 0.3 to 0.1 MW/m
2
by
retarding the DI SI from 30 to 90 "CA, which is very limited. With even higher EGR at 30%, misfire
occurs for tests with DI SI between 75 and 115 0CA.
20
40
100-
60
80
100
120
140
160
180
100
-
10
U
uEGR=1O%
EGR=O0%
*EGR=20%
*EGR=25%
0 0.1
--
20
40
60
0.
0.
EGR=30%
SO
100
120
140
DI timing (*CA ABDC Intake)
160
160
Figure 5-43 KI of tests with intake temperature 100 0C, CR 15, RPM 1200, MAP 1.8 bar
This is saying the window for EGR dilution adjustment is narrow to make both stable combustion
and mitigated engine knock. Notice tests with DI SOI later than 135 *CA are only conducted with 30%
EGR, this is to demonstrate the trend for the test set without severely vibrating the engine by low EGR
and high MPRR tests.
A full map of MPRR versus EGR sweep is shown below in Figure 5-44. The MPRR of tests with
different EGR dilution rates barely changes with DI SI being retarded from 30 *CA to 60 "CA. With DI
SI later than 60 *CA, tests with EGR=0, 10%, and 20% have MPRR increasing monopoly, indicating
advanced ignition preparation processes for these tests, which might be caused by the high cylinder
temperature induced faster evaporation. For tests with EGR rate 25%, MPRR is kept almost constant for
DI SI through out 30 "CA to 90 *CA, and after that, it climbs up as other curves.
147
20
so
40
so
100
140
120
ISO
160
EGR=O%
EGR=10%
EGR=20%
- EGR=25%
* EGR=30%
-aa
40-
-.
40
30
30U
*
-- 'U
a",
20-
20
10 -
10
20
40
60
60
100
180
140
120
10
Di timing (*CA ABDC intake)
Figure 5-44 MPRR of tests with intake temperature 100 *C, CR 15, RPM 1200, MAP 1.8 bar
Figure 5-45 shows the CA10 and CA50 of above tests. For EGR rate of 0, 10%, and 20%, the
combustion phase isn't retarded by the stratified fuel but rather advanced by local rich zones from SI
as early as 60 'CA. Tests with EGR 25% makes a difference and retards both the CA10 and CA50 for 2 *CA
by varying the S01 from 30 *CA to 90 *CA, although this wouldn't do much to the burn duration as shown
in Figure 5-46.
40
20
100
80
60
120
140
160
180
190
20
40
60
80
100
190
*
U
U
U
184-
1864
162-
182
a
a
-a
20
180
M
EGR=O%
a EGR=10%
-a- EGR=20%
" EGR=25%
-" EGR=30%
176.
U
190
186
0
V0
178-
I 80
195
160
EGR=O%
EGR=10%
-aEGR=20%
- - EGR=25%
- EGR=30%
-186
E
140
-**
166
188-
120
40
60
--
180
*~*NJ
U,
-180
178
I
120
140
100
80
Di timing (*CA ABDC intake)
176
180
174
180
Ito M
20
I
40
1
140
120
100
80
0
DI timing ( CA ABDC Intake)
60
Figure 5-45 Combustion Phase with intake temperature 100 *C
148
175
160
Is0
20
40
60
80
100
120
140
-5 --
o
*
160
180
EGR=O%
EGR=10%
EGR=20%
- EGR=25%
EGR=30%
5
4U
20
4
60
80
100
120
DI timing (*CA ABDC
Figure 5-46
140
160
180
Intake)
Burn duration with intake temperature 100 *C
In summary, intake temperature significantly impacts the CAI combustion. In this test we found:
1)
High intake temperature worsens the engine knock tendency by diminishing fuel stratification
effect, as it pushes the in-cylinder charger into a relatively more temperature sensitive regime
(i.e. less fuel stratification sensitive) for fuel ignition.
2)
High intake temperature also increases the in-cylinder equivalence ratio, although the charge
density is also reduced. These changes impact ignition.
3)
Fuel stratification effects are more prominent both for knock reduction with DI SCII in the middle
of compression stroke, and for pressure rise enhancement for late DI SCII at around 20 "CA
before TDC.
5.2.5
Engine Speed Impact
Engine speed has significant impacts on the fuel mixing process and engine timings. With higher
engine speed, the turbulence inside engine cylinder could be enhanced, yet the absolute time for fuel to
evaporate, mix, and chemically prepare for ignition would be less. To understand the impact of engine
speed, a set of comparison tests was conducted with all the other conditions kept the same as base case
but the engine speed at 1500 rpm.
1.
Knock Performance
149
Figure 5-47 shows the KI comparison between 1200 rpm and 1500 rpm tests with no EGR.
It can be
seen that there is little difference in the KI for both speeds in the early DI SOI regime
(30 to 130 *CA) and
in the very late DI SC0 regime (170 to 190 *CA).The KI values at SI of 130 and 140
CA, however, are
much higher for the 1500 rpm than the 1200 rpm case.
20
40
60
80
100
120
140
160
180
100
100
-U-
RPM1200 EGR=O%
RPM1500 EGR=O%
10--
10
0 1
0.01 1
20
200
40
60
01
80
100
120
140
180
DI timing ("CA ABDC Intake)
180
0.01
200
Figure 5-47 KI of tests for engine speed impact study
Note that data were not collected for tests with DI SCl between 140 "CA and 170 *CA, because of
too severe engine knock and data collection was skipped for the purpose of protecting the cylinder
pressure transducer from thermal shock.
Figure 5-48 shows the MPRR of above tests. At 1500 rpm, MPRR is reduced for tests with early DI
SI (30 "CA to 120 *CA) by more than 5 MPa/ms. It starts ramping up drastically in the higher engine
speed tests and finally gives a worse knock performance for tests with DI SI later than 130 *CA.
Much
increased MPRR with DI SOl between 140 *CA and 170 *CA made it impossible to record the cylinder
pressure data. After 170 *CA, the KI drops with retarded DI SI, while very moderate improvement
seen
for higher engine speed tests.
150
20
40 I
40
*
*
60
80
100
120
140
160
180
200
40
RPM1200 EGR=O%
RPM1500 EGR=O%
-
30-
30
U
- 20
20 -
U
m
10 -
10
U-
u'm
20
40
80
80
100
120
140
160
180
-0
200
DI timing (*CA ABDC Intake)
Figure 5-48 MPRR of tests for engine speed impact study
2.
Combustion Performance
As mentioned above, the effect of retarded combustion phase due to shortened absolute time for
everything, and enhanced fuel stratification effect due to less time for mixing might simultaneously
impact the combustion and knock performance to make differences between tests with different engine
speeds. To understand the combustion phase effect and fuel stratification effect separately, analysis to
combustion is conducted.
CA10 and CA50 for above tests are presented in Figure 5-49.
20
40
60
80
100
120
140
160
180
12 .
190 U-
U-
190
200 -
186
186 -
184 -
-'
U-
40
80
100
120
140
160
200
195-
uI.
E
184
195
190
190--
182
182 -
*
U.
-U-U
200
205
180
-U-RPM1200 EGR=O%
RPM1500 EGR=O%
-U-
185-
185
180-
180
180
RPM12 )0 EGR=O%
-RPM15 10 EGR=0%
178
178 -
176
20
60
188
U,
4
S
20
205-
U
'a'
188 -
200
192
40
80
80
100
120
140
180
180
175
200
20
40
80
60
100
120
140
160
DI timing (*CA ABDC Intake)
DI timing (*CA ABDC Intake)
Figure 5-49 Combustion phase of tests for engine speed impact study
151
180
175
200
It is clear that with an increment of 300 rpm in engine speed, CA10 gets retarded for 4 "CA in the
early DI timing tests and as much as 5 "CA in the tests with DI SOI in the middle of compression stroke,
and CA50 gets retarded for 5 *CA in the early DI timing tests and as much as 7 "CA in the tests with DI
SOI in the middle of compression stroke. For late DI SOl tests, the combustion phase is barely retarded
0
by higher engine speed (- 1 CA). Therefore the combustion phasing effect in early and very late DI SOI
regimes definitely impacts.
see that
Figure 5-50 shows the burn duration comparison between different engine speeds. Easy to
a generic extension of burn duration for around 2 *CA shows up in the early DI SOI tests by higher engine
to the middle of
speed, and the extension is further enhanced by retarding the DI SOI further
burn duration
compression stroke, reaching over 5 *CA at DI SOI 120 *CA. With DI SOl later than that, the
is
shows no significant difference between the two engine speeds. So the fuel stratification effect
enhanced as well, seeing from the mid-compression stroke DI SOI difference.
20
40
so
so
100
120
140
160
180
20
- RPM1200 EGR=%
URPM1 500 EGR=0%
20-
200
158
15
10-10
10
0
/
0
20
40
80
160
140
120
100
80
DI timing (*CA ABDC Intake)
180
200
Figure 5-50 Combustion phase of tests for engine speed impact study
To summarize, higher engine speed shortens the absolute time for many events of the CAl
combustion, therefore:
1)
The combustion phase could be effectively retarded as the fuel evaporation, mixing, and
the
chemical preparation processes for ignition get retarded, this should generally reduce
engine knock tendency. However the effect could be different in different DI SOI regimes.
152
2)
The fuel stratification effect is enhanced in all aspects, as with less time to mix, the knock
mitigation effects can be strengthened with DI SOI in mid-compression stroke, as well we the
knock enhancement effects with DI SOI near TDC.
5.3
Conclusion
CAI combustion was conducted with Haltermann 437 gasoline on an original Renault 1.9 L diesel
engine in a single firing cylinder under intake boost conditions. The effect of fuel stratification on CAI
knock mitigation was systematically studied by varying the DI fuel portion and timing, while the EGR
effect was also studied orthogonally. Tests have been conducted on two different compression ratios
and achieved very different results. The operational parameters are varied later on the low CR engine
set up to study the impact of fuel split ratio, intake temperature, and engine speed on the effectiveness
of both effects.
On the high CR (19) engine, both single DI and PFS tests show improvement in knock performance
comparing with PFI tests by having DI SOI in the early compression stroke, however varying the DI timing
didn't further improve it. With late injection near TDC, the combustion becomes very severe and heavy
knocks show up. For very late DI SOI, the combustion in single pulse DI tests becomes unstable very
soon with the further retarded Dl SOI, while the PFS tests show a stable two-combustion-event
phenomenon with the ignition timing earlier than the DI timing, however with increasingly large energy
loss. Seeing the amplitude of knock governed by the amount of DI fuel, we see fuel split ratio can be
used to effectively adjust the fuel stratification extent. In general, the results didn't show much potential
for using fuel stratification as effective tool on CAI knock mitigation. EGR effectively reduces the incylinder temperature, thereby retards the combustion phasing, and reduces the fast heat release rate of
the fuel, contributing to the CAI engine mitigation. The effectiveness of EGR is not equal for all the tests.
On the low CR (15) engine, a base case test set was first conducted showing much improved knock
mitigation effect by placing the DI SOI in the middle of compression stroke, while its effectiveness shows
dependence on in-cylinder temperature by comparison among tests with difference EGR. It is found fuel
stratification impacts the combustion phase so that the knock tendency, while it also has direct impact
on burn duration. EGR shows positive impacts on knock mitigation; however misfire is also induced with
too heavy EGR in certain cases. The behavior of CAI combustion with late DI near TDC and very late Dl
shows very alike trend as on the high CR engine.
153
The impact of fuel split ratio is investigated by comparing tests with same operation conditions but
different PFI/DI ratio (40/60, 60/40, and 80/20). It is found that DI fuel enhances the fuel stratification
effects in all perspective: knock mitigation for mid-compression stroke DI SOI, and drastic pressure rise
for late DI SOI, as well as misfire for very late DI SOI.
The impact of intake temperature is investigated by comparing tests with same operation conditions
but different intake temperatures (60 0C, 80 "C, and 100 "C). In the conducted tests, high intake
temperature worsens the engine knock tendency by diminishing fuel stratification effect via the
different fuel ignition sensitivity in different temperature regimes, as well as by counteracting the EGR
effect on temperature curbing. Low intake temperature enhances the fuel stratification effects both for
knock reduction with DI SOI in the middle of compression stroke, and for pressure rise enhancement
with late DI SOI near TDC.
The impact of engine speed is investigated by comparing tests with same operation conditions but
different rpm values (1200 and 1500). It was found the combustion phase could be effectively retarded
by higher engine speed and generally reduce the engine knock tendency. On the other hand the fuel
stratification effect is enhanced in all aspects, as with less time to mix, the knock mitigation effects can
be strengthened with DI SOI in mid-compression stroke, as well we the knock enhancement effects with
DI SOI near TDC.
In summary, the experiment in this research shows that fuel stratification and EGR have potential on
CAI knock mitigation under boosted condition. However to make it effective, subtle design of the
combustion system is needed to optimize the engine performance. That includes considerations in many
aspects such as in-cylinder temperature range (CR, intake temperature), EGR range, DI timing, DI
portion, engine speed range, and so on. Also lots of these factors have substantial interdependence. This
research can be used as reference for such work.
154
Chapter 6. Effect of Fuel Stratification on CAI Combustion
Chapter 5 reported the knock performance and corresponding combustion analysis of a CAI engine
with DI SOI varied for fuel stratification creation and adjustment. From the results we see the effect of
fuel stratification could be very different with DI SOI in different regimes, and under different operation
conditions. The effectiveness of fuel stratification could be significantly impacted by these factors as
well. To understand the fundamental mechanism of the engine performance, analysis is conducted
regarding the fuel stratification, its impact on combustion phasing and burn duration, fuel ignition
sensitivity to in-cylinder conditions, and sequential combustion phenomenon. This chapter links the
engine behavior with the fuel ignition characteristic, and the fuel stratification adjustment.
6.1
Stratification and Its Impact to CA10
6.1.1
The Existence of Fuel Stratification with DI SOI
The combustion equivalence ratio could significantly impact HC and NO emissions. Therefore the HC
and NO emissions could be used to indicate the in-cylinder equivalence ratio status. For premixed
combustion, equivalence ratio is fairly evenly distributed. For stratified fuel, the fuel-air mixture parcel
with different equivalence ratio burn within once engine cycle. Parcels with close to stoichiometric
combustion would have high NO and low HC emissions. By examining the exhaust NO and HC, we should
be able to profile the fuel stratification status at ignition.
Due to the limited emissions measurement, emissions data were not recorded for a typical test set
as described in Chapter 5. Below we look at two sets of HC & NO versus DI SOI on CR 19 and CR 15
engines with different PFI/DI ratios respectively for the purpose of conceptually understanding the
evolution of fuel stratification with DI SOI variation.
Figure 6-1 shows the HC and NO emissions measured for a test set on the CR 19 engine, with
PFI/DI= 30/70, intake temperature 80 'C, intake pressure 1.9 bar, and engine speed 1200 rpm. The EGR
rates swept includes 0, 30%, and 40%. Clearly for DI SOI in the early compression stroke (prior to 120
'CA), the NO emission is negligible, while HC stays above 400ppm, showing a globally lean combustion
characteristic, corresponding to the global equivalence ratio at around 0.25 (EGR=0). With DI SOI
retarded to 120 'CA, the HC emission increases slightly; at EGR=40%, it hits 600 ppm. This is saying that
155
combustion occurred with locally even leaner mixture due to the incomplete mixing before ignition,
meaning with stratification enhanced by later DI SOI, the fuel mass has a broader distribution against
equivalence ratio, and the leaner mixture dominated the ignition timing (CA10), as majority fuel is with
relatively lower equivalence ratio. Notice by having 30% of fuel delivered via PFI, a background
equivalence ratio of 0.075 could be created before the DI fuel delivered, and with not long enough
mixing time, regions with
4) closed to that value could still exist and ignite once the in-cylinder
temperature is elevated after ignition.
Retarding DI SOI further, the HC emissions starts to drop down while the NO emission picks up. This
means for tests with DI SO in this regime, the combustion occurs at a moment that rich zones are more
dominating. HC emissions reach the valley point at around 200 ppm for all three curves with different
EGR rate when DI SOI gets to 150 *CA, where NO emission increases with DI SOI retarding with an even
steeper slope. This means the majority fuel burns with a higher equivalence ratio, and very limited
amount of locally lean zones still give certain amount of HC generation. By the time DI SO[ retards to 160
0
CA, the NO emission peaks, as bulk part of the fuel doesn't mix well before ignition, inducing high
equivalence ratio combustion. Clearly the HC level was kept relatively constant comparing with earlier,
indicating a relatively constant amount of fuel burned lean.
After 160 'CA, the NO starts to drop down and HC picks up a little bit as DI SOI keeps retarding. The
decreasing of NO emission is mainly a temperature effect, as the minimum mixing time for the fuel to hit
the ignition "rich limit" requires certain amount of time (~15 'CA here). With, DI SOI at 160 'CA, the
process could be finished before TDC and the combustion just burst out right before TDC. However with
even later DI SOI, the "rich limit" couldn't be hit until after TDC, while the cylinder temperature starts
dropping down due to the downward piston motion, which makes it even longer to chemically prepare
for the ignition and a retarded combustion phase starts to be. This will induce lower combustion
temperature, and therefore less NO emissions. On the other hand, the slight increase of HC emission is
very likely because the later DI fuel ignites as a bulk first, while the lean zones formed with the premixed
PFI fuel wouldn't have time to mix with the bulk part DI fuel so burns shortly after the major ignition is
set off with the low equivalence ratio kept. The trend got enhanced with even later DI SOI.
It looks EGR could differentiate the NO emissions quite significant in the "rich-ignition" SOI regime
(150 to 180 0CA) while HC is moderately impacted. This might be because NO emissions are more
sensitive to cylinder temperature.
156
1200
-4-HC EGR=O
-U-HC EGR=30%
'I
800
0
z
EGR1 =40%
,"-. L~U~=Q7
-X-NO EGR =0
-A-NO EGR =30%
- -NO EGR= 40%
'
1000
600
400
I
200
0
140
120
100
80
180
160
D1 Sol (CA ABDC intake)
Figure 6-1 HC and NO emissions with DI SOI variation for tests with PFI/DI=30/70, Tin = 80 *C, Pin = 1.8
bar, engine speed 1200 rpm, and total fuel 23 mg/cycle on CR19 engine
A similar group data obtain from a set of tests on CR 15 engine is shown in Figure 6-2.
2350
C EGR= 0 %
HC EGR = 15%
HC EGR = 30%
1950
NO EGR = 0%
-I
1550
NO EGR =15%
N0 EGR = 30%
CL
o
z 1150
I
750
350
-50
o
20
40
60
80
100
120
140
160
180
Di So1 (deg CA ABDC [intake])
Figure 6-2 HC and NO emissions with Dl SOI variation for tests with PFI/DI=40/60, Tin = 80 *C, Pin = 1.8
bar, engine speed 1200 rpm, and total fuel 23 mg/cycle on CR 15 engine
157
The tests were done with PFI/DI ratio at 40/60, intake temperature 80 0C, intake pressure 1.8 bar,
engine speed 1200 rpm, and total fuel 23 mg/cycle. The EGR rates were swept for 0, 15%, and 30%. This
group of curves shows a similar pattern as curves in
Figure 6-1. However, since the CR 15 engine
provides a more fuel stratification sensitive environment, the data for EGR =15% is of void for tests with
DI SOI in the middle of compression stroke due to misfire caused by fuel stratification induced
combustion phasing retard. The EGR 30% data is missed in the entire lean ignition regime (DI SOt earlier
than 130 *CA). With very late DI SOI after 160 0CA, unstable combustion occurs, and with higher EGR
rate, it occurs at earlier DI SOI.
From above we see that with DI SOI retarded, fuel stratification shows up on both CR 19 and CR 15
engines with PFS. However the evolution of it presents very different status of fuel stratification: for
very early DI SO (~30 to 90
0
CA), the charge ignites when the majority fuel reaches the global
equivalence ratio in a near premixed manner, as the time available for mixing before ignition is more
than sufficient; with a bit later DI SOI (~90 to 120 0 CA), the ignition sets off before all the fuel reaches
the global equivalence ratio, with a big part of fuel burns at leaner conditions, and this provides
potential for retarded combustion phase and extended burn durations if the fuel ignition is sensitive to
equivalence ratio distribution; DI SO even later (~120 to 160 'CA), ignition occurs when majority fuel
has a higher equivalence ratio than the global one, yet over the "rich limit" for ignition, and this makes
the combustion phasing advanced and burn duration shortened; with even later DI SOI (later than 160
'CA), the hard limit for fuel to mix and reach the ignition "rich limit" pushes the ignition timing after
TDC, with a decreasing in-cylinder temperature due to the piston motion, combustion phasing gets
retarded and burn duration extended.
The evolution of fuel distribution inside engine cylinder is driven by the DI timing and DI portion, but
the impact of it to combustion and knock performance is also dependent on the chemical preparation
process for ignition and combustion. The sensitivity of ignition to fuel distribution is also playing a role.
6.1.2
A conceptual Model of Fuel Stratification on CA10
Liven-good Wu approach [53] is a classic method for estimating the ignition timing of fuel autoignition on internal combustion engines. The basic idea is that the with the in-cylinder charger
compressed by piston, the ignition delay ignition shortens with the increased temperature and charge
density, and the eventual ignition timing depends on the integration of the ignition delay on the
(temperature, charge density) path. Once the integration reaches 1, the fast heat release will occur
158
immediately. In experimental tests, we normally approximate that moment with CA10, the moment
when 10% of the total energy is released.
For premixed charge, the status of in-cylinder gas parcels are ideally homogeneous, and the ignition
delay integral of individual parcel is the same as the global integration. Then the CA10 represents the
start point of the fast heat release for all the charge parcels. However for fuel-stratified charge, the
integration of ignition delay for charge parcels with different equivalence ratio can be different,
meaning the actually "CA10" for different parcels are not necessarily the same. The measured "CA10"
depends on the fuel distribution it represents the moment where 10% of the total energy released by
the first fraction of the charge parcels that ignite. If there is a distribution of fuel mass against ignition
delay, the CA10 will be determined by the ignition delay of the fuel masses. Since the ignition delay is
sensitive to the equivalence ratio, CA10 would be determined by the equivalence ratio distribution.
Therefore, to see how fuel stratification impacts CA10, we could analyze the possible fuel mass
distribution against equivalence ratio, and trace the potential Liven-good Wu integration for fuel mass
with different equivalence ratio, and compare the moment that the majority fuel mass burst into
combustion. So the link between fuel stratification and CA10 variation is made by fuel mass distribution
against equivalence ratio (eventually ignition delay).
The time between CA10 and DI SOI may be interpreted as the total ignition delay in the engine. As
shown in Figure 6-3 is the variation of "CA10-DI SOI" with DI SOI for the single pulse injection tests as
described in Chapter 5. We can see that with DI SOI keeps retarding, the distance between CA10 and SOI
keeps shortening in the earlier DI SOI regime, as proper mixing could always be achieved before TDC to
facilitate the ignition, however when it comes to later than 160 'CA, the minimum mixing to make the
fuel ignitable with very limited oxygen could not be finished before TDC, and extra time is needed to
make that happen. Notice the entire fuel injection process takes up to 4.36 'CA for single pulse DI tests
and around only 2.8 'CA for tests with PFI/DI= 60/40 in the CR15 base case tests, which are short
compared to other ignition preparation events (e.g. mixing, chemical preparation, etc.)
159
EGR = 0
-1--EGR = 20%
-- 6-- EGR = 30%
"EGR = 40%
22
160
+EGR = 0
N EGR = 20%
EGR = 30%
X EGR = 40%
140
120
20
N
X
18
100V
5
-so
60
16
14
4
20
70
20
120
12
150
170
160
170
180
DI S01 (deg CA ABDC [intake])
DI SO1 (deg CA ABDC [intake])
Figure 6-3 CA10-DI SOI for single pulse DI test in Chapter 5
This is saying that for the DI fuel to be ignited at all, proper mixing is needed, so for early DI S01
tests, the actually chemical preparation wouldn't start until this process is finished. Although the ignition
for PFS tests with sufficient PFI fuel are not limited by this factor as the PFI part of fuel could ignited first,
the physical preparation for the DI part of fuel is still experienced before it enters into the chemical
preparation process.
Besides that, the chemical process is also limited by temperature as well. From RCM tests and some
CHEMKIN Modeling work (will be shown later) we see that the ignition wouldn't occur for the fuel under
use in a constant volume combustion set up until the temperature reaches 650K, no matter if the charge
density is large enough, meaning the ignition delay of fuel mixture is "infinitely long" in temperature
under 650K. Notice the mixing process still goes on after the Liven-good Wu integration is started.
Take the CR 15 engine base case test (PFS, PFI/DI= 60/40, 1200rpm) as an example. Figure 6-4 shows
the "ignition delay" defined as the distance between CA10 and DI SOI, and the compression
temperature and pressure as in the tests with EGR= 15%. Easy to see the in-cylinder temperature
reaches 650K at around 140 *CA, meaning that's the point for mixture which is over the "rich limit" to
start the Liven-good Wu integration.
For tests with DI SO in regime A, relatively long time is available for mixing before the cylinder
temperature reaches 650 K. The test with DI SO at point P will have fairly homogeneous mixture by 140
160
'CA, and the majority fuel mass will have equivalence ratio at around the global value 0.4, so the
integration starts and completes in a manner closed to totally premixed charge.
When it comes to point Q, the equivalence ratio distribution is varied due to shorter time for DI fuel
to mix. The majority fuel mass is of a lower equivalence ratio as reflected from Figure 6-1 Figure and 62, since quite a bit mixture might still be under the background equivalence ratio of 0.24 created by the
PFI fuel, although some parcels might of a higher value. The Liven-good Wu integration for different
parcels here would start with different slope from 140 *CA on, and would eventually reaches to different
values by 184 'CA, where the measured CA10 of test with DI SOI at point P locates. These values added
up by mass weight distribution wouldn't reach 1, as compared to point P test, the majority fuel mass is
with equivalence ratio in the leaner side since 140 *CA. The integration will have to keep on until 190
'CA, where 10% of the total energy release is completed, and this is a later CA10 comparing with the
very early DI SOI tests.
So or CAIO
CA deg ABDC-intake
20
0
40
60
80
100 120
140
160 180
200
0I
Ignition delay
50
a
Soi
100 150-
F
B
A
<200
1000 -
-
-
-
R
G
80
-
r ,-
8004 c
40-
T
S600-a
20
R
4Q
20
40
60
S T
0
80 100 120 140 160 180 200
CA deg abdc-intake
Figure 6-4 CA10, compression temperature and pressure with DI SOI for CR15 base case test
161
For test with DI SOI even later at point F, the mixing process needed for mixture to reach the "rich
limit" will penetrate into time period after 140 'CA, and once that's finished, the Liven-good Wu
integration would be initiated immediately for parcels with much higher equivalence ratio. That gives a
much faster integration process, and the ignition would be reached earlier. By very late DI SOI, the
ignition of very rich parcels could be much earlier than the leaner parcels, and the heat release of the
rich parcels could accelerate the temperature raise inside engine cylinder and push the leaner ones to
complete the integration faster, and finally gives an advanced CA10 comparing to tests with a bit earlier
DI SOL. Figure 6-1 and Figure 6-2 could reflect the existence of both lean and rich parcels in this regime
as well.
This trend peaks at DI SOI 160 'CA, point R as shown in Figure 6-4. After 160 0CA, the minimum time
needed for DI fuel to mix and enter into the proper equivalence regime makes the start of Liven-good
Wu integration of the rich parcels after TDC, and the effect of temperature reduction caused by
downward piston motion kicks in. The ignition delay of the mixture parcels could be globally extended,
and retarded CA10 is resulted, as in point G shown in Figure 6-4. With DI SOI later than that, the PFI fuel
starts the Liven-good Wu integration before the DI fuel could finish the "rich limit" mixing, and by the
time rich parcels starts on the integration track, the charge temperature is already pretty low, canceling
out the shortening ignition delay effect of rich zones. With even retarded DI SOI later than 185 0 CA, as in
point T, the Liven-good Wu integration is completed by PFI fuel solely before the DI SOI, although the
fast heat release process of the 60% PFI fuel is set off, the DI fuel still needs some more time to mix and
enter into the combustible regime, so the ignition of DI fuel will eventually presents as a separate event
from the PFI fuel ignition, giving a much relaxed heat release process. The constant retarding of CA10 for
tests with CA10 later than DI SOI could be due to the reduced wall temperature from the weaker
combustion of these tests.
6.2
Charge Temperature, Stratification, and Heat Release.
As described in Chapter3, the burn duration on CAI engine should be largely governed by the
ignition delay differences of fuel parcels at ignition. Because once the fast heat release process is set off,
there will be no material difference among the fast heat release process of different fuel or mixtures.
The fuel stratification and temperature stratification inside engine cylinder both contribute to this
differentiation, as shown in the theoretical analysis in Chapter 3, concisely indicated by Equation [3.28].
However we found this is not totally the truth. Strong connection is found between the fuel fast heat
162
release rate and the engine heat release process in this research. The fuel fast heat release rate is largely
dependent on the charge temperature at ignition, and this is the link between CA10 and the engine heat
release process.
6.2.1
Combustion Phasing, Charge Temperature, and Heat Release
MPRR represents the heat release intensity of CAI combustion, as seen from analysis in Chapter 3,
indicated by Equation [3.12].
Figure 6-5 shows the MPRR variation with CA10 for the CR 15 base case
tests. We can see that the data is almost monotonic, showing a strong correlation between the
combustion phasing and heat release process. This is a temperature effect due to the different fast heat
release rate of the mixture under different ignition temperature. To illustrate further, the MPRR values
are plotted against the unburned gas temperature at CA10 calculated by a cycle simulation as well.
40
*EGR=O
* EGR=10%
AEGR = 15%
X EGR =20%
35
30
E 25
40
EGR
35
AO%A
30
*10%
r 25
2,20
015%4
0.20
S
A XX
I0.15
20
5
al
15
4L
C.10
020%
crp *A
*
5
0
176
178
180
182
1
186
1
n
190
850
CAM ( deg CA ABD (intake])
860
870
880
890
900
910
920
Charge temperature @ CAI0 (K)
Figure 6-5 MPRR vs. CA10 and Temperature at CA10 for CR15 base case tests
A set of RCM tests was conducted to support the hypothesis. By operating the RCM with Halterman
437 gasoline at cD = 0.4 and fixed compressed charge density of 7 kmol/m 3, the ignition delay and the
rate of pressure rise are measured for a range of compression temperature between 744K and 915K,
covering the temperature at CA10 for the engine experiment between 860 K and 910 K as shown in
Figure 6-4. The charge is diluted with nitrogen of 15% by mole to mimic the 15% EGR. Different
compression ratios (12, 18 and 20) and initial temperatures are used to get the compression
temperatures, and the corresponding compression pressure is between 43 and 53 bar.
163
The maximum pressure rise rate, which is an indicator of the heat release rate, from the RCM data is
shown in Figure 6-6. The rate increases almost linearly with the temperature. The engine data in Figure
6-5, however, have shown more sensitivity to temperature. One possible reason could be that the RCM
experiments were conducted at a fixed compressed density, while the density of the charge at
combustion changes with combustion phasing.
12
10
CL
4
2
0
930
880
830
780
730
Temperature (K)
Figure 6-6 MPRR of mixture ignited at CA10 temperatures of CR15 base case tests
The MPRR dependence on temperature observed in the engine experiments is therefore consistent
with the finite heat release rate that increases with temperature observed in the RCM experiment.
6.2.2
Stratification and Burn Duration
However looking at the data in Figure 6-5 closely, we find that there are "branches" in both the very
high and very low MPRR regime, showing different heat release characteristic for test with the same
duration
CA10 or ignition temperature. This is caused by fuel stratification. Figure 6-7 shows the burn
see that
variation with CA10 for tests in Figure 6-5, giving a better idea of the phenomenon. Easy to
difference
different burn durations are shown for tests with the same EGR rate at the same CA10. The
in the
between tests with very late DI SOI and early SOI is pretty obvious, saying that the heat release
because the
very late DI SOl tests are much more severe than the early injection tests. This is
high
combustion occurs in the very late DI SOI tests when lots of mixture parcels are still with pretty
stroke is
equivalence ratio. The difference among tests with DI SOl around the middle of compression
with DI SOI
not very discernable, yet could be recognized in a zoom-in graph with only data for tests
around mid-compression stroke shown as in Figure 6-8.
164
20
is
16
tko
14
20-0-EGR = 0
-M-EGR = 10%
,r-EGR =15%
-XEGR=-20%
12
Very late Dl SOI
10
8
6
4
Eary D1 SO[
2
176
180
178
182
184
188
186
190
CA10 (deg CA ABDC [intake])
Figure 6-7 Burn Duration vs. CA10 of CR15 base case tests
11
10
9
--
8
-ir-EGR =15%
EGR=O
-U-EGR =10%
Later DI SOl
7
6
E S
4
Earlier D1 Mo
3
2
180
182
184
186
188
190
CA10 (deg CA ABDC [intake])
Figure 6-8 Burn Duration vs. CA10 of CR15 base case tests zoom-in
165
Slight difference of MPRR in tests with DI SOI before and after the peak point of burn duration in the
early injection regime indicates the sequential burn effect induced by stratified charge.
As mentioned in Chapter 5, the effect of stratification could be universally magnified with lower
PFI/DI ratio. Figure 6-9 shows the MPRR versus CA10 for tests with all the conditions the same as CR 15
base case tests but PFI/DI=40/60 rather than 40/60. The DI SOI range of this test set is: for EGR=0 and
EGR = 10%, DI SOI is from 30 to 170 0CA; for EGR = 15%, DI SOI is from 30 to 160 'CA; and for EGR =30%,
DI SOI is from 130 to 165 *CA. Clearly both the sequential burn effect by mid-compression stroke DI SI
and the rich zone effect by late DI SOI around 160 *CA are enhanced comparing with the base case test
set.
40
Rich Zone Effect
'n
E 30
C.
Se uential Burn Effect
=
-MEGR =0%
-- EGR =15%
- 4-EGR = 30%
10
0
176
178
180
182
184
186
1
190
CAIO (CA ABDC [intake])
Figure 6-9 MPRR vs. CA10 for stratification comparing test set of CR 15 base case with lower PFI/DI
ratio at 40/60
Therefore the heat release process is mainly governed by combustion phasing via the ignition
temperature, as the fast heat release rate upon different ignition temperature actually impacts the
general heat release rate. While fuel stratification impacts the combustion phasing with the differently
distributed equivalence ratio, it could also directly impact the heat release process. From the test results
166
in this research we can see that the effect of rich zones is very obvious and sequential burn effect
requires relatively stronger stratification to be presented.
6.3
Sensitivity of Ignition to Fuel Stratification
From Chapter 5 we found that even fuel stratification exists for tests with both CR19 and CR15, the
effect on combustion phasing retarding and burn duration extending only shows up in the CR 15 tests. It
is hypothesized that this is because the fuel ignition sensitivity to the equivalence ratio distribution after
compression is relatively strong in the CR 15 engine cylinder than in the CR 19 engine cylinder. To
validate that, a set of simulation work was conducted to understand the in-cylinder conditions in the
beginning and the end of the compression before ignition for above tests. This is realized by
propositioning the in-cylinder (temperature, charge density) trajectory along compression on an ignition
delay contour map with temperature and charge density as coordinate, inspired by the Liven-good Wu
approach for auto-ignition. To focus on the sensitivity study, the mixture is assumed homogeneous for a
unified equivalence ratio on the map. The ignition delay contour map is organized with ignition delay
data obtained from a set of CHEMKING [57] modeling, and the in-cylinder process is modeled by an
quasi-dimensional model developed at MIT for the performance and emissions prediction of
reciprocating SI engine cycles [58].
6.3.1
Ignition Delay Data
A set of simulation work was conducted with CHEMKIN to obtain a database of the ignition delay
values needed for the proposition. A constant volume combustion model was set up to simulate the
RCM experiments for the ignition delay calculation. Equivalence ratio was fixed at 0.4 and 0.25 to
simulate the global equivalence ratios in the tests on CR 15 and CR19 engines respectively as reported in
Chapter 5. No inert gas dilution was used. The charge density was swept from 0.1 Kmol/m
Kmol/m
3
3
to 1.4
for all the tests to cover the data range from very early compression to immediately before
ignition, and ignition temperature was swept from 650 K to 1100 K, to cover the entire possible ignition
temperature regime. The composition of the fuel was simulated by setting it with iso-octane and nheptane blends of corresponding mass fraction.
Before the model was run for the target test sets, a small group of simulation was conducted with
the same conditions as the RCM tests in Chapter 4 to validate the model set up. The experimental
167
3
condition is constant charge density at 0.4 Kmol/m , with EGR=0, and 4)=1.0. Figure 6-10 shows the
comparison of ignition delay data from simulation and experiment. A good consistency can be observed.
100
* Chemkin Results
* Experiment Results
=Al
E
U
10
C
R
1
1.5
1.4
1.3
1000/T (1/K)
1.2
Figure 6-10 CHEMKIN Model validation with experiments
Figure 6-11 and Figure 6-12 show the ignition delay curves obtained from the modeling work for
equivalence ratio 0.25 and 0.4 respectively.
4 -=0.25
"0""C=.lKmoI/m3
E
V
C0.2 KmoI/m3
-- 0--C=0.3 Kmol/Im3
100
-0""C=0.4 KmoI/m3
*""C=O.5 KmoI/m3
Kmol/m3
O s-C=0.7
KmoI/m3
C=0.3 KmoI/m3
C=0.9 Kmol/m3
C=1.2 KmoI/m3
W--e"C=0.6
10
C
C
""E"C=1.1KmoI/m3
1
"C-1.3
KmoI/m3
C--1.4 Kiol/m3
0
0.1
0.85
0.95
1.05
N
0
S
0-
S
1.15
1.25
1.35
1.45
1.55
1000/T (1/K)
Figure 6-11 CHEMKIN Model results for (D=0.25
168
U,4
(P
vI
E
U
W
-"""""C=
1KmoI/m3
-0mC=0.2 Kmol/m3
-- 46"C=0.3 Kmol/m3
100
"**C=0.4 Kmol/m3
-""C=0.5 Kmol/m3
"'"C=0.6Kmol/m3
C=0 7Kmol/m3
i"
C=0.8 Kmol/m3
C=0.9 Kmol/m3
C=1.0 Kmol/m3
"""C=1.1Kmol/m3
C=1.2 KmoI/m3
C=1.3 Kmol/m3
C=1.4 Kmol/m3
10
C
1
tko
0.1
0.85
0.95
1.05
1.15
1.35
1.25
1.45
1.55
1000/T (1/K)
Figure 6-12 CHEMKIN Model results for D=0.4
With above data, a contour of constant ignition delay can be achieved on a (temperature, charge
density) map for each equivalence ratio.
6.3.2
Compression Process in the Engine Cycle
A Fortran program was developed at MIT [58] for dimensional cycle simulation of the operating
cycle of reciprocating SI engines. The program calculates temperature and pressure in the combustion
chamber as well as performance and emissions as a function of operating conditions and engine design.
By putting in the engine geometry, fuel property, and other operating parameters from experiments, a
"temperature vs. charge density" trace can be achieved for an engine cycle.
In this research,
experiments in comparable conditions with those reported in Chapter 5 is used as reference for the incylinder compression process simulation.
Figure 6-13 shows the pressure trace comparison between the simulation and experiment for the
CR19 test. The test run is with no EGR, intake pressure 1.9 bar, intake temperature 80 'C, and engine
speed 1200 rpm. The fuel is delivered by PFI portal at 70 *CA before TDC of intake at an amount of
16 mg/cycle, giving an equivalence ratio of 0.25. Easy to see that the experiment and simulation
result aligns well in most part. The difference in the peak pressure could be from heat transfer
inconsistences. However since the concerned part for this research is the compression process
before ignition, which looks well taken care of by the model.
169
180
Experiment Result
160
-
140
i--Simulaton Result
120
100
S80
I60
40
20
0
100
300
250
200
150
Crank Angle (ABDC [intake)
Figure 6-13 Pressure trace comparison for simulation and experiment for test with CR19, PF fuel=16
mg/cycle, Tin= 80 *C, Pin=1.9 bar, engine speed 1200 rpm, no EGR
Figure 6-14 shows the pressure trace comparison for the CR15 test. The test run is with no EGR,
intake pressure 1.8 bar, intake temperature 80 *C, and engine speed 1200 rpm. The fuel is delivered
by PFI portal at 70 'CA before TDC of intake at an amount of 23 mg/cycle, giving an equivalence
ratio of 0.4. Good consistency in the compression part can be seen.
160
,'Experiment
140
-Simulation
Result
Result
120
100
'80
60
40
20
0
100
150
200
250
300
Crank Angle (ABDC [intake])
Figure 6-14 Pressure trace comparison for simulation and experiment for test with CR15, PFI fuel=23
mg/cycle, Tin= 80 *C, Pin=1.8 bar, engine speed 1200 rpm, no EGR
170
With above results, the in-cylinder temperature-charge density traces are propositioned with the
ignition contour map as in Figure 6-15 and Figure 6-16 for the CR19 and 15 tests separately. The ignition
delay value is shown on the contour curves, and the distance between curves in the horizontal direction
indicates the sensitivity of ignition delay to in-cylinder temperature, while the distance between
curves
in the vertical direction indicates the sensitivity of ignition delay to charge density, in other words, the
fuel density. The sensitivity reflected by charge density with homogenous mixture could be interpreted
as the sensitivity to fuel concentration distribution, or equivalence ratio distribution with stratified fuel
in cylinder.
A10 (100% 1)
I
=930.44 K
n 1.1779 KmoI/m
=4.3705 ms
1/lgnition Delay Contour (1/ns)
3
80% I
1.2
00
60% I
E
0
V
E
40% I
0
a
20%1I
0.6
10% I
0 0.4
0
0.2
650
%I
T = 665.49 K
n=0.3734 KmoI/m
=95.6640 ms
.x
700
750
800
850
900
Temperature (K)
950
'?o
1000
%
1050
1100
Figure 6-15 In-cylinder conditions along compression process for CR 19 engine
Also marked in the figure are points indicating the percentage of the Liven-good Wu integration
completed. Easy to see that points at later time has larger significance to the ignition timing, as the
majority portion of the integration is contributed by these parts.
By comparing the sensitivity of ignition delay to charge density in the near-integration- completion
area, we could see the potential effectiveness of fuel stratification to ignition timing (CA10) adjustment.
171
3
By comparing the sensitivity at point CA10, we could see the potential effectiveness of fuel stratification
to sequential burn enhancement.
1/lgnition Delay Contour (1/ms)
-
CA10 (100%1)
T= 853.59 K
n= 0.8802 KmoI/m
,= 8.1047 ms
3
b
1.2
60% I
K
50%1
\o..
E 1
0
E
0.8
0.6
40%1
30%1
-..
aR
0
20%1
*0~,PIP
C
10%1
0.4
0.0
%1 I
T= 659.08 K
.........
n= 0.3454
0.2
650
KmoI/M
,r= 73.9519 ms
700
750
800
900
850
Temperature (K)
950
1000
1050
1100
Figure 6-16 In-cylinder conditions along compression process for CR 15 engine
In Figure 6-15, near 60% of the chemical preparation process for ignition is completed in the
temperature regime over 900K, where the ignition delay mostly relies on the temperature rather than
charge density or equivalence ratio. In Figure 6-16, the last 60% of the integration is completed in the
temperature regime from around 825 K to 860 K, where the ignition delay relies on charge density much
heavier. Clearly, CR19 engine has relatively high temperature before and at ignition that are over the
NTC regime, therefore the ignition delay has low sensitivity to fuel stratification but high sensitivity to
temperature. In the other hand, CR15 engine has relatively low temperature before and at ignition that
are right around NTC regime, although it's a bit on the edge of the higher temperature side as well. This
makes the effect of combustion phasing retarding and sequential burn by fuel stratification on the CR19
engine not show up at all, but appears with optimized operation conditions on the CR15 engine.
172
3
The RCM tests reported in 6.2.1 also provide support from the experimental perspective for this
argument with its data on ignition delay. Figure 6-17 is the ignition delay curve for mixture with charge
density 0.7 Kmol/m 3 , equivalence ratio 0.4, and inert gas dilution rate 15%, simulating the global charge
status of the CR 15 base case test with 15% EGR. Clearly, the temperature before and at ignition as
mentioned above is right around NTC region for such conditions, although a temperature reduction
effect by 15% EGR should also be counted in.
100
E
C
0
10
930
905
880
855
830
805
780
755
730
Temperature (K)
Figure 6-17 Ignition delay curve for CR 15 base case test with 15% EGR
Also easy to see from both Figure 6-15 and Figure 6-16 that since the temperature by CA10 is
relatively high, the differentiation of ignition delay caused by fuel stratification is naturally not as strong
as in lower temperature regime, this is saying that the effect on ignition timing control by using fuel
stratification is born to be stronger than its impact to sequential burn, when fast heat release of certain
parcels already set off in a fuel stratified cylinder. This explains the much stronger effect of fuel
stratification on combustion phasing retarding than sequential burn effect. The situation might be
different if the ignition on engine can be achieved at very low temperature that's before or around NTC
regime of the fuel mixture. This might be possible for very low speed operations, in which the Livengood Wu integration could take time to finish before the temperature gets too high by compression.
173
6.4
Conclusion
1.
Fuel Stratification created by varying DI SOI timing is the result of the competition between
the fuel and air mixing process and the chemical ignition preparation process. The
stratification extent at ignition can be reflected by HC and NO emissions measured from the
exhaust, and could be used to infer the stratification status before ignition.
2.
Fuel stratification could effectively impact the engine ignition timing (CA10). This is because
by having a distribution of fuel mass against the equivalence ratio, the chemical preparation
process for ignition of different mixture parcels could be well differentiated, and the
measured CA10 on engine is mostly determined by the mass averaged low-temperature heat
release processes. The competition among equivalence ratio and fuel mixing process for DI
SOI in different regimes is illustrated with a conceptual model.
3.
Opposing to common wisdom, the fast heat release governed by ignition temperature does
impact the engine burn duration. The ignition temperature is largely determined by ignition
timing (CA10).
4.
Fuel stratification directly impacts the burn duration on engine in two different manners
depending on the stratification extent: rich zone effect and sequential burn effect.
5.
In-cylinder conditions impact the sensitivity of ignition delay to fuel stratification significantly,
and this determines the effectiveness of fuel stratification for both combustion phasing
retarding and sequential burn.
6.
That combustion phasing retarding effects showing a more significant impact than sequential
burn could be explained by the different temperature of in-cylinder charge before and after
ignition.
174
Chapter 7. Summary and Conclusion
This research tries to explore the potential of using fuel stratification and EGR on CAI engine knock
mitigation under intake boost conditions. With success on that, the high load limit of CAI engines could
be then expanded into a range comparable to conventional automotive engines, and help push the
environment and fuel economy friendly CAI combustion technology into industrial practice. To do that, a
series research work was conducted and reported in this document, with efforts primarily via
experiments and also certain necessary simulation. The work is based on the preliminary understanding
of the CAI engine knock physics and the potential of fuel stratification and EGR from research work done
by researchers at MIT and other organizations. The major steps of the entire research include:
1.
Fuel characteristics investigation: obtaining the fuel auto-ignition characteristics under different
concentration, dilution, and temperature corresponding to in-cylinder conditions of typical high
load CAI combustion.
2.
Engine performance tests: conducting tests on CAI engine with different operation conditions
for varied fuel stratification and EGR dilution, demonstrating the effectiveness of the
techniques.
3.
Analysis and Interpretation: connecting the fundamentals with the engine performance for a
comprehensive and thorough understanding of the effectiveness of fuel stratification and EGR
under different operation conditions for CAI engine knock mitigation.
This chapter summarizes the key findings of above work and draws conclusions of the research
project.
7.1
CAI Engine Knock and Potential Solutions
CAI engine knock is caused by abrupt in-cylinder pressure rise that induces pressure wave
development inside engine cylinder. The pressure oscillation of the gas charge causes engine block to
resonate and lead to acoustic wave radiation perceived as unacceptable noise. With the majority of the
energy contained in the relatively low frequency (4-8KHz) vibrations, the intensity of the CAI engine
knock could be measured and calculated from the pressure oscillations, as in Equation [3.10].
175
Knock Intensity, defined as in Equation [3.12] was adopted as metric of the CAl engine knock. MPRR
was found to be effective in indicating the CAI engine knock propensity as a parameter obtained from
the low pass pressure trace measured in the CAI combustion process. This bridges up the knock
performance and combustion process on CAI engines, laying the foundation of controlling engine knock
by adjusting combustion system. The inconsistence between MPRR and KI under certain conditions
needs to be noticed and addressed for correlating.
Fuel stratification was proposed as a promising strategy for CAI knocks mitigation based on
theoretical analysis, while it was also pointed out that the effectiveness of it could be impacted by incylinder conditions and engine operation parameters. EGR was proposed as another technique for CAI
knock control considering its dilution effects on in-cylinder temperature and fuel fast heat release
process.
7.2
Fuel Ignition Characteristics
Ignition delay and MPRR (notice the difference between this term and MPRR in engine operations)
were identified as measurements to profile the fuel ignition characteristics. Experiments were
conducted on RCM to map out these two parameters for the fuel used in this research, Haltermann 437
gasoline, under different dilution composition (air, nitrogen, and simulated complete combustion
products), dilution rates, and charge densities in the interested ignition temperature regime (700 to
1000 K). Two sub sets of experimental tests were conducted with constant fuel density and constant
charge density separately to approach the fundamentals for fuel stratification effect and EGR effect on
CAI combustion.
With constant charge density but differentiated equivalence ratio, the sensitivity of ignition delay to
temperature and equivalence ratio distributions was studied. It was found that in the NTC regime, the
ignition is more equivalence ratio sensitive, and in higher or lower temperature regime, the temperature
dominates the ignition delay of the fuel mixture. This means to make fuel stratification effective on
extending the heat release process on CAI engine, the in-cylinder conditions need to be adjusted into
certain temperature regime to best utilize the stratification effect.
With constant fuel density, fixed power output was simulated for high load operation of CAI
combustion with different levels of dilution. The effect of air dilution and inert gas dilution was obtained
and compared with each other. The data shows that inert gas dilution effectively extends the ignition
176
delay, and also relaxes the fast heat release process in certain temperature regime, while extra oxygen
provided by air dilution might shortens the ignition delay with a much weaker power. Comparison test
sets conducted with simulated combustion products including CO 2 and H20 vapor as diluents confirmed
finding, and shows that in reality the oxygen in the EGR gas could reduce the inert gas dilution effect
slightly. Tests with higher fuel density was also conducted for the purpose of more comprehensive
understanding of the topic, showing that higher fuel density significantly shortens the ignition delay and
enhances the MPRR for the same ignition temperature. This provides extra support to the in-cylinder
process analysis as both charge density and fuel density could change with varied combustion phasing.
7.3
Engine Performance with Fuel Stratification and EGR
CAI combustion was operated with the same fuel on a modified diesel engine in a single firing
cylinder under intake boost conditions. The effect of fuel stratification and EGR was systematically
studied by varying the DI fuel portion and timing, while the EGR effect was also swept independently.
The engine compression ratio, intake temperature, and engine speed were varied to study the
effectiveness of both strategies under different operation conditions.
With a high compression ratio at 19, neither single pulse DI nor PFS tests show obvious effect on
knock performance improvement by directly injecting DI fuel in the early compression stroke. With late
DI injection around 20 *CA before TDC, the combustion becomes violent and engine knocks severely. For
very late DI SOI, single pulse DI tests gives unstable combustion with the further retarded DI SOI, while
the PFS tests show a stable two-combustion-event phenomenon with the CA10 earlier than DI SOl.
Although the knock tendency of the engine with dual injection and very late DI timing is slightly reduced,
energy loss was also found due to the too late combustion. EGR was found effectively reduces the incylinder temperature, thereby retards the combustion phasing, and reduces the fast heat release rate of
the fuel, contributing to the CAI engine mitigation.
With a low compression ratio at 15, a base case test set with PFI/DI ratio of 60/40 was conducted
and showed much improved knock performance by placing the DI SOI in the middle of the compression
stroke. The effectiveness of the strategy shows dependence on EGR rate. It was found with different
levels of fuel stratification, both the combustion phase and the burn duration could be impacted. EGR
shows positive impacts on knock mitigation, however misfire is also observed with too heavy EGR in
177
certain tests. The behavior of CAI combustion with late DI near TDC and very late DI shows very alike
trend as on the high CR engine.
The impact of fuel split ratio is investigated by comparing tests with same operation conditions but
different PFI/DI ratio (40/60, 60/40, and 80/20). It is found that DI fuel enhances the fuel stratification
effects in all perspective: knock mitigation for mid-compression stroke DI SOI, and drastic pressure rise
for late DI SOI, as well as misfire for very late DI SO.
The impact of intake temperature is investigated by comparing tests with same operation conditions
but different intake temperatures (60 "C, 80 "C, and 100 "C). In the conducted tests, high intake
temperature worsens the engine knock tendency by diminishing fuel stratification effect via the
different fuel ignition sensitivity in different temperature regimes, as well as by counteracting the EGR
effect on temperature curbing. Low intake temperature enhances the fuel stratification effects both for
knock reduction with DI SOI in the middle of compression stroke, and for pressure rise enhancement
with late DI SOI near TDC.
The impact of engine speed is investigated by comparing tests with same operation conditions but
different rpm values (1200 and 1500). It was found the combustion phase could be effectively retarded
by higher engine speed and generally reduce the engine knock tendency. On the other hand the fuel
stratification effect is enhanced in all aspects, as with less time to mix, the knock mitigation effects can
be strengthened with DI SOI in mid-compression stroke, as well we the knock enhancement effects with
DI SOI near TDC.
The engine experiments demonstrated the potential of fuel stratification and EGR for CAI knock
mitigation while also showed their limitations and possible problems. The engine configuration and
operation parameters impact the engine responses significantly and differently.
7.4
The Effect of Fuel Stratification and EGR
Analysis and supplemental experiment and simulation work was done to interpret the engine
performance with the fundamentals. It was found fuel stratification impacts CAI combustion via both
combustion phasing and burn duration adjustment, and the EGR effect is realized through both cylinder
temperature modification, and fast heat release rate reduction. Other operation parameters such as
compression ratio, intake temperature, engine speed and etc. could all influence the final engine
178
performance direct or indirectly. A diagram as shown in Figure 7-1 might be used to outline the links
between the knock performance and operation conditions on a CAI engine.
Compression Ratio
Charge Density
Rich/lean limit
for Ignition
)Intake
_________
--------
-
--
-
--
Charge Temperature
During Compression
Fast Heat Release
Ignition Temperature
Burn Duration
Extent of Fuel
Stratification @
Tww
Intake Pressure
7!j________
Temperatur
Cylinder Pressure
Sequential
ti__ Bur
__
650KJ
Knock Intensity
Ignition Sensitivity
to Fuel Stratification
PFI/DI Ratio
Extent of Fuel
Stratification @
CA10
Engine Speed
Figure 7-1 Links between CAI engine knock performance and operation conditions
More specifically, it was found:
1.
Fuel Stratification created by varying DI SOl timing is the result of the competition between
the fuel and air mixing process and the chemical ignition preparation process. The
stratification extent at ignition can be reflected by HC and NO emissions measured from the
exhaust, and could be used to infer the stratification status before ignition.
2.
Fuel stratification could effectively impact the engine ignition timing (CA10). This is because
by having a distribution of fuel mass against the equivalence ratio, the chemical preparation
process for ignition of different mixture parcels could be well differentiated, and the
measured CA10 on engine is mostly determined by the mass averaged low-temperature heat
179
release processes. The competition among equivalence ratio and fuel mixing process for DI
SOI in different regimes is illustrated with a conceptual model.
3.
Opposing to common wisdom, the fast heat release governed by ignition temperature does
impact the engine burn duration. The ignition temperature is largely determined by ignition
timing (CA10).
4.
Fuel stratification directly impacts the burn duration on engine in two different manners
depending on the stratification extent: rich zone effect and sequential burn effect.
5.
In-cylinder conditions impact the sensitivity of ignition delay to fuel stratification significantly,
and this determines the effectiveness of fuel stratification for both combustion phasing
retarding and sequential burn.
6.
That combustion phasing retarding effects shown a more significant impact than sequential
burn could be explained by the different temperature of in-cylinder charge before and after
ignition.
7.5
Conclusion
The research work shows that fuel stratification and EGR strategies could be utilized for CAI knock
mitigation and high load limit expansion. The fuel stratification could be used for combustion phasing
adjustment as well as sequential burn enhancement, while EGR is effective for in-cylinder temperature
modification and fuel fast heat release relaxation. To make these strategies sufficiently effective while
not draw in extra set backs (e.g. misfire), a full picture of the fuel characteristics, fuel stratification
extent, and engine operating parameters is needed for the design and configuring of the CAI combustion
system. Outlining the potential links of the knock performance and engine fundamentals, this research
could serve as a reference for the future development of CAI engines with the capability of knock free
high load operations.
180
Appendix: Specifications of Haltermann 437 Gasoline
HALTERMANN
PRODUCT:
PRODUCT CODE:
TEST
EPA TIER 11 EEE
FEDERAL REGISTER
HF437
METHOD
Distillation - iBP
ASTM D86
5%
10%
20%
30%
40%
50%
60%
70%
80%
90%
95%
Distillation - EP
Recovery
Residue
Loss
Gravity
ASTM D4052
Density
ASTM D4052
Reid Vapor Pressure
ASTM D323
Reid Vapor Pressure
ASTM D5191
Carbon
ASTM D3343
Carbon
ASTM E191
Hydrogen
ASTM E191
Hydrogen/Carbon ratio
ASTM E191
Oxygen
ASTM D4815
Sulfur
ASTM D5453
Lead
ASTM D3237
Phosphorous
ASTM D3231
Composition, aromatics
ASTM D1319
Composition, olefins
ASTM D1319
Composition, saturates
ASTM D1319
Benzene
ASTM D3606
Particulate matter
ASTM D5452
Oxidation Stability
ASTM D525
Copper Corrosion
ASTM D130
Gum content, washed
ASTM D381
Fuel Economy Numerator/C Density ASTM E191
C Factor
ASTM E191
Research Octane Number
ASTM D2699
Motor Octane Number
ASTM D2700
Sensitivity
Net Heating Value, btullb
ASTM D3338
Net Heating Value, btu/lb
ASTM D240
Color
VISUAL
I
S
-F
-F
75
HALTERMANN S
M WN
TARGET MAX
95
75
95
"F
*F
120
135
120
135
200
230
200
230
305
325
305
325
UNITS
FED
MIN
*F
*F
*F
*F
-F
-F
*F
'F
'F
415
415
vol%
vol %
vol %
API
58.7
kgl
psI
8.7
psi
wt fraction
wt fraction
wt fraction
mole/mole
wt %
wt%
0.0015
g/gal
g/gal
vol %
vol %
vol %
vol%
mg/l
minutes
Report
Report
Report
61.2
9.2
58.7
0.734
8.7
61.2
0.744
9.2
Report
Report
Report
Report
Report
0.0080
0.05
0.005
35.0
10.0
0.05
0.0035
0.01
0.005
35.0
10.0
0.0025
Report
Report
1
240
1
5
2441
mg1l00mis
2401
Report
93.0
96.0
7.5
7.5
Report
btu/lb
btuIb
Report
Report
Report
I
DOW RESTRICTED
181
- For
intemal use only
TYPICAL
RESULTS
89
117
131
154
181
209
224
234
244
267
321
335
402
98.0
1.0
1.0
58.9
0.743
9.1
9.00
0.8664
0.8641
0.1309
1.805
<0.05
0.0029
<0.01
<0.0008
30.7
0.5
68.8
0.1
0.6
>1000
1
<1
2433
0.9992
97.4
89.0
8.4
18450
18435
CLEAR
(this page intentionally left blank)
182
Reference
1.
Amir Gamal Maria, On Fuel Selection in Controlled Auto-Ignition Engines: The Link between Intake
Conditions, Chemical Kinetics, and Stratification, Ph.D. Thesis, Dept. of Mech. Eng., MIT, 2012.
2.
Onishi, S., Jo, S., Shoda, K., Jo, P. et al., "Active Thermo-Atmosphere Combustion (ATAC) - A New
Combustion Process for Internal Combustion Engines," SAE Technical Paper 790501, 1979.
3.
Fuquan Zhao, Thomas N. Asmus, Dennis N. Assanis, John E. Dec, James A. Eng, Paul M. Najt,
Homogeneous Charge Compression Ignition (HCCI) Engines, SAE International, 2003.
4.
Andreae, M., Cheng, W., Kenney, T., and Yang, J., "On HCCI Engine Knock," SAE Technical Paper
2007-01-1858, 2007.
5.
Heywood, J.B., Internal Combustion Engine Fundamentals, McGraw-Hill, 1988.
6.
Tian, G., Wang, Z., Ge, Q., Wang, J. et al., "Mode Switch of SI-HCCI Combustion on a GDI Engine,"
SAE Technical Paper 2007-01-0195, 2007.
7.
Sakashita, Y., Suzuki, H., and Takagi, Y., "Controlling Onset of Heat Release by Assisted Spark Ignition
in Hydrogen HCCI Engine Supported by DME Supplement," SAE Technical Paper 2009-01-1419, 2009.
8.
Lawler, B., Ortiz-Soto, E., Gupta, R., Peng, H. et al., "Hybrid Electric Vehicle Powertrain and Control
Strategy Optimization to Maximize the Synergy with a Gasoline HCCI Engine," SAE Int. J. Engines
4(1):1115-1126, 2011, doi:10.4271/2011-01-0888.
9.
Chang, K., Lavoie, G., Babajimopoulos, A., Filipi, Z. et al., "Control of a Multi-Cylinder HCCI Engine
During Transient Operation by Modulating Residual Gas Fraction to Compensate for Wall
Temperature Effects," SAE Technical Paper 2007-01-0204, 2007.
10. Scaringe, R., Wildman, C., and Cheng, W., "On the High Load Limit of Boosted Gasoline HCCI Engine
Operating in NVO Mode," SAE Int. J. Engines 3(1):35-45, 2010, doi:10.4271/2010-01-0162.
11. Sj6berg, M. and Dec, J., "Influence of Fuel Autoignition Reactivity on the High-Load Limits of HCCI
Engines," SAE Int. J. Engines 1(1):39-58, 2009, doi:10.4271/2008-01-0054.
12. Hildingsson, L., Johansson, B., Kalghatgi, G., and Harrison, A., "Some Effects of Fuel Autoignition
Quality and Volatility in Premixed Compression Ignition Engines," SAE Int. J. Engines 3(1):440-460,
2010, doi:10.4271/2010-01-0607.
13. Turner, D., Tian, G., Xu, H., Wyszynski, M. et al., "An Experimental Study of Dieseline Combustion in
a Direct Injection Engine," SAE Technical Paper 2009-01-1101, 2009.
14. Szybist, J., Edwards, K., Foster, M., Confer, K. et al., "Characterization of Engine Control Authority on
HCCI Combustion as the High Load Limit is Approached," SAE Int. J. Engines 6(1):553-568, 2013,
doi:10.4271/2013-01-1665.
15. Yun, H., Wermuth, N., and Najt, P., "High Load HCCI Operation Using Different Valving Strategies in a
Naturally-Aspirated Gasoline HCCI Engine", SAE Int. J. Engines 4(1):1190-1201, 2011,
doi:10.4271/2011-01-0899.
183
16. Attard, W. and Parsons, P., "A Normally Aspirated Spark Initiated Combustion System Capable of
High Load, High Efficiency and Near Zero NOx Emissions in a Modern Vehicle Powertrain," SAE Int. J.
Engines 3(2):269-287, 2010, doi:10.4271/2010-01-2196.
17. Szybist, J., Nafziger, E., and Weall, A., "Load Expansion of Stoichiometric HCCI Using Spark Assist and
Hydraulic Valve Actuation," SAE Int. J. Engines 3(2):244-258, 2010, doi:10.4271/2010-01-2172.
18. Polovina, D., McKenna, D., Wheeler, J., Sterniak, J. et al., "Steady-State Combustion Development of
a Downsized Multi-Cylinder Engine with Range Extended HCCI/SACI Capability," SAE Int. J. Engines
6(1):504-519, 2013, doi:10.4271/2013-01-1655.
19. Curran, S., Hanson, R., and Wagner, R., "Effect of E85 on RCCI Performance and Emissions on a
Multi-Cylinder Light-Duty Diesel Engine," SAE Technical Paper 2012-01-0376, 2012.
20. Dec, J. and Yang, Y., "Boosted HCCI for High Power without Engine Knock and with Ultra-Low NOx
Emissions - using Conventional Gasoline," SAE Int. J. Engines 3(1):750-767, 2010, doi:10.4271/201001-1086.
21. Dec, J., Yang, Y., and Dronniou, N., "Improving Efficiency and Using E10 for Higher Loads in Boosted
HCCI Engines," SAE Int. J. Engines 5(3):1009-1032, 2012, doi:10.4271/2012-01-1107.
22. Christensen, M. and Johansson, B., "Supercharged Homogeneous Charge Compression Ignition
(HCCI) with Exhaust Gas Recirculation and Pilot Fuel," SAE Technical Paper 2000-01-1835, 2000.
23. Olsson, J., Tunestil, P., Haraldsson, G., and Johansson, B., "A Turbo Charged Dual Fuel HCCI Engine,"
SAE Technical Paper 2001-01-1896, 2001.
24. Bessonette, P., Schleyer, C., Duffy, K., Hardy, W. et al., "Effects of Fuel Property Changes on HeavyDuty HCCI Combustion," SAE Technical Paper 2007-01-0191, 2007.
25. Christensen, M., Johansson, B., Amneus, P., and Mauss, F., "Supercharged Homogeneous Charge
Compression Ignition," SAE Technical Paper 980787, 1998.
26. Johansson, T., Johansson, B., Tunest5l, P., and Aulin, H., "HCCI Operating Range in a Turbo-charged
Multi Cylinder Engine with VVT and Spray-Guided DI," SAE Technical Paper 2009-01-0494, 2009.
27. Yelvington, P.E., and Green, W.H., "Prediction of the Knock Limit and Viable Operating Range for a
Homogeneous-Charge Compression-Ignition (HCCI) Engine", SAE Paper 2003-01-1092, 2003.
28. Eng, J., "Characterization of Pressure Waves in HCCI Combustion," SAE Technical Paper 2002-012859, 2002.
29. Santoso, H., Matthews, J., and Cheng, W., "Characteristics of HCCI Engine Operating in the NegativeValve-Overlap Mode," SAE Technical Paper 2005-01-2133, 2005.
30. Andreae, M., Cheng, W., Kenney, T., and Yang, J., "Effect of Air Temperature and Humidity on
Gasoline HCCI Operating in the Negative-Valve-Overlap Mode," SAE Technical Paper 2007-01-0221,
2007.
31. Cairns, A. and Blaxill, H., "The Effects of Combined Internal and External Exhaust Gas Recirculation
on Gasoline Controlled Auto-Ignition," SAE Technical Paper 2005-01-0133, 2005.
184
32. Sjoberg, M. and Dec, J., "EGR and Intake Boost for Managing HCCI Low-Temperature Heat Release
over Wide Ranges of Engine Speed," SAE Technical Paper 2007-01-0051, 2007.
33. Sj6berg, M., Dec, J., and Hwang, W., "Thermodynamic and Chemical Effects of EGR and Its
Constituents on HCCI Autoignition," SAE Technical Paper 2007-01-0207, 2007.
34. Wildman, C., High Load Limits of the Controlled Autoignition Engine, Ph.D. Thesis, Dept. of Mech.
Eng., MIT, 2009.
35. Scaringe, R., Extension of the High Load Limit in the Homogenous Charge Compression Ignition
Engine, Ph.D. Thesis, Dept. of Mech. Eng., MIT, 2009.
36. Noda, T. and Foster, D., "A Numerical Study to Control Combustion Duration of Hydrogen-Fueled
HCCI by Using Multi-Zone Chemical Kinetics Simulation," SAE Technical Paper 2001-01-0250, 2001.
37. Sj6berg, M., and Dec, J., "Comparing Enhanced Natural Thermal Stratification against Retarded
Combustion Phasing for Smooting of HCCI Heat-Release Rates", SAE Paper 2004-01-2994, 2004.
38. Kakuho, A., Nagamine, M., Amenomori, Y., Urushihara, T. et al., "In-Cylinder Temperature
Distribution Measurement and Its Application to HCCI Combustion," SAE Technical Paper 2006-011202, 2006.
39. Kuboyama, T., Moriyoshi, Y., Hatamura, K., Suzuki, M. et al., "Effect of Fuel and Thermal
Stratifications on the Operational Range of an HCCI Gasoline Engine Using the Blow-Down Super
Charge System," SAE Int. J. Engines 3(1):666-680, 2010, doi:10.4271/2010-01-0845.
40. Sj6berg, M., and Dec, J., "Smooting HCCI Heat-Release Rates Using Partial Fuel Stratification with
Two-Stage Ignition Fuels", SAE Paper 2006-01-0629, 2006.
41. Dec, J.E., Yang, Y., and Dronniou, N., "Boosted HCCI - Controlling Pressure-Rise Rates for
Performance Improvements using Partial Fuel Stratification with Conventional Gasoline", SAE Paper
2011-01-0897, 2011.
42. Johansson, T., Johansson, B., Tunestel, P., and Aulin, H., "The Effect of Intake Temperature in a
Turbocharged Multi Cylinder Engine operating in HCCI mode," SAE Int. J. Engines 2(2):452-466, 2010,
doi:10.4271/2009-24-0060.
43. Widd, A., Johansson, R., Borgqvist, P., Tunest51, P. et al., "Investigating Mode Switch from SI to HCCI
using Early Intake Valve Closing and Negative Valve Overlap," SAE Technical Paper 2011-01-1775,
2011.
44. Hanson, R., Kokjohn, S., Splitter, D., and Reitz, R., "An Experimental Investigation of Fuel Reactivity
Controlled PCCI Combustion in a Heavy-Duty Engine," SAE Int. J. Engines 3(1):700-716, 2010,
doi:10.4271/2010-01-0864.
45. Gaurav Mittal, A Rapid Compression Machine - Design, Characterization, and Autoignition
Investigations, Ph.D. Thesis, Dept. of Mech. and Aero. Eng., Case Western Reserve University, 2006.
46. Amir Gamal Maria, Rapid Compression Machine Manual, Sloan Automotive Lab, MIT, 2010.
47. Mitsubishi, Engine F9Q Series, Mitsubishi Motor Corporation, 2000.
185
48. Kurz Instruments, Series 504FT In-Line Mass Flow Meter Transmitter User's Guide, Kurz Instruments,
Inc., 2003.
49. Maria, A., Cheng, W., Kar, K., and Cannella, W., "Understanding Knock Metric for Controlled AutoIgnition Engines," SAE Int. J. Engines 6(1):533-540, 2013, doi:10.4271/2013-01-1658.
50. C. S. Draper, "Pressure Waves Accompanying Detonation in the Internal Combustion Engine",
Journal of the Aeronautical Sciences (Institute of the Aeronautical Sciences), Vol. 5, No. 6, pp. 219226,1938.
51. Wildman, C., Scaringe, R., and Cheng, W., "On the Maximum Pressure Rise Rate in Boosted HCCI
Operation," SAE Technical Paper 2009-01-2727, 2009.
52. Wai K. Cheng, "Impact of temperature profile and temperature sensitivity of ignition delay on HCCI
heat release rate", Personal note, Mar. 18, 2008.
53. Livengood, J.C., Wu, P.C., "Correlation of Autoignition Phenomena in Internal Combustion Engines
and Rapid Compression Machines", Fifth Symposium (International) on Combustion, 347-56, 1955.
54. Tairin Hahn, Ignition Study in Rapid Compression Machine Thesis for the Degree of Mechanical
Engineer, Dept. of Mech. Eng., MIT, 2009.
55. K. Fieweger, R. Blumenthal, G. Adomeit, "Self-ignition of S.I. engine model fuels: A shock tube
investigation at high pressure", Comb. Flame 109:599-619, 1997.
56. X. He, M.T. Donovan, B.T. Zigler, T.R. Palmer, S.M. Walton, M.S. Wooldridge, A. Atreya, "An
experimental and modeling study of iso-octane ignition delay times under homogeneous charge
compression ignition conditions", Combustion and Flame 142: 266-275, 2005.
57. Reaction Design, CHEMKIN, Reaction Design, http://www.reactiondesign.com/products/chemkin/.
58. Stephen Gregory Poulos, The effect of combustion chamber geometry on S.I. engine combustion
rates: a modeling study M.S. Thesis, Dept. of Mech. Eng., MIT, 2009.
186