EFFECTS OF OPERATING CONDITIONS,
COMPRESSION RATIO, AND GASOLINE REFORMATE
ON SI ENGINE KNOCK LIMITS
by
Michael D. Gerty
B.A.Sc., Mechanical Engineering
University of British Columbia, 2001
Submitted to the Department of Mechanical Engineering
in Partial Fulfillment of the Requirements for the Degree of
Master of Science in Mechanical Engineering
at the
MASSACHUSETTS INS
OF TECHNOLOGY
Massachusetts Institute of Technology
JUN 16 2005
June 2005
© 2005 Massachusetts Institute of Technology
All Rights Reserved
LIBRARIES
Signature redacted
Signature of the Author ............................................................................... .
Department of Mechanical Engineering
May 13,2005
Signature redacted
Certified by ........................................................................ J~~ 'B"li~~;~d
Sun Jae Professor of Mechanical Engineering
Thesis Advisor
Signature redacted
Accepted by ....................................................... ~.......... , .................. : ...... .
Lallit Anand
Professor, Department of Mechanical Engineering
Chairman, Department of Graduate Commitee
E.
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EFFECTS OF OPERATING CONDITIONS,
COMPRESSION RATIO, AND GASOLINE REFORMATE
ON SI ENGINE KNOCK LIMITS
by
Michael D. Gerty
Submitted to the Department of Mechanical Engineering
on May 13, 2005 in Partial Fulfillment of the
Requirements for the Degree of Master of Science in
Mechanical Engineering
ABSTRACT
A set of experiments was performed to investigate the effects of air-fuel ratio, inlet boost
pressure, hydrogen rich fuel reformate, and compression ratio on engine knock behavior. For
each condition the effect of spark timing on torque output was measured. Knock limited spark
advance was then found for a range of octane number (ON) for each of three fuel types; primary
reference fuels (PRFs), toluene reference fuels (TRFs), and test gasolines.
A new combustion phasing parameter based on the timing of 50% mass fraction burned,
termed "combustion retard", was found to correlate well to engine performance. Increasing airfuel ratio increases the combustion retard required to just avoid knock for PRFs and has little
effect for TRFs. Combustion retard also increases more with inlet pressure and decreases more
with reformate addition for PRFs than for TRFs. Both fuel types responded similarly to
increased compression ratio. The trends for gasoline are about halfway between PRFs and TRFs.
Experiments were also performed to determine the response of mid-load indicated
efficiency to air-fuel ratio, load, and compression ratio. At a compression ratio of 9.8:1, relative
net efficiency improvement is about 2.5% per unit compression ratio. Efficiency peaks at about
14:1 with a maximum benefit of 6-7%.
Detailed chemical kinetics were combined with a cylinder pressure based end-gas
modeling methodology to successfully predicted the response of PRFs to compression ratio and
air-fuel ratio, and the response of TRFs to boost. The difference between the response of PRFs
and TRFs to air-fuel ratio was also captured. Constant volume chemistry modeling found that
hydrogen slows alkane autoignition reactions by consuming hydroxy radicals in the end gas.
Reforming 30% of the fuel entering an engine decreases the required fuel quality 10 ON
or more, which would allow increased compression ratio or increased turbocharging without
increasing combustion retard. A simplified analysis indicates that increasing compression ratio
and downsizing the engine to maintain constant maximum torque would increase fuel efficiency
by about 9%. Turbocharging and downsizing would increase fuel efficiency by about 16%.
Thesis Advisor: John B. Heywood
Title: Sun Jae Professor of Mechanical Engineering
3
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ACKNOWLEDGEMENTS
I would like to thank Professor John Heywood for allowing me the invaluable
opportunity to focus on a subject that I enjoy at one of the worlds greatest engineering
institutions. It has been an honor to be mentored by a leader of such great character,
knowledge, and integrity. His respect and high professional expectations motivated me
to excel in my areas of strength, to develop new skills, and to strive for greater
achievement. He was a key part of making my time at MIT a rewarding experience.
This work would not have been possible without my predecessor Jennifer Topinka. Her
creative and yet practical ideas pushed the plasmatron project into the new territory that I
lucky to be able to explore. My valued co-workers and office mates Joshua Goldwitz and
Ziga Ivanic made coming to the lab an adventure, and helped make my transition into
MIT and the Sloan Automotive Laboratory much smoother and more enjoyable. Thanks
also to Ferran Ayala and Bridget Revier, my teammates who will be carrying the
plasmatron/knock torch when I leave.
The Sloan Automotive Laboratory has been a great place to work, learn, and socialize
with fellow students and dedicated MIT employees. It's impossible to name everybody,
but I'd like to thank those that made life at the lab interesting and helped me with my
research. Among the employees, I'd like to thank Karla Stryker for keeping Professor
Heywood's chaotic schedule under control; Thane Dewitt and Raymond Phan for
facilitating virtually all experimental work that goes on in the lab; and Professor Jim
Keck, Professor Wai Cheng, and Dr. Jim Cowart for their useful insight into the
complexities of this project.
This project wouldn't exist without its sponsor Arvin Meritor with Rudy Smaling leading
the way. The Plasma Science and Fusion Center also deserves credit for developing a
key element of this research, the plasmatron fuel reformer.
I sincerely appreciate all of the support that I received from my home country, Canada.
Dr. Robert Evans at UBC, and my colleagues and supervisors at Westport Innovations
afforded me the opportunity to gain experience in the engine business, which paved the
way to my application and acceptance to MIT and the Sloan Lab. A scholarship from the
Natural Sciences and Engineering Research Council of Canada helped to support me
financially. The company of my fellow ex-pats, Jeff Jocsak, JP Urbanski, and Devon
Manz, allowed me to enjoy some of the finer points of Canadian culture while away from
home. Also, absence has not diminished my appreciation for Perry Wong, Kelly
Cameron, Graham Knutson, and all my other friends from beautiful British Columbia.
My family - Mom (Carol), Dad (Brian), Krista, and Tyler - have always been there for
support and encouragement when I needed it most.
Finally, my fiancee Phoi has been my most inspirational force, even from thousands of
kilometers away. I'd like to express my sincere gratitude for the loving support,
encouragement, and patience that she provided me every step of the way.
5
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6
TABLE OF CONTENTS
A bstract.................................................................................
- - -. 3
........ .. ---................--
Acknowledgements .......................................................................................
. ---------------...............
L ist of Figures.........................................................................
....
Chapter 1. Introduction..................................................................................
1.1 Knock in Spark Ignition Engines ........................................................................
1..1 A utoignitionChem istry........................................................................................
1.1.2 FactorsAffecting Autoignition in SI Engines....................................................
1.2 Efficiency Effects of Compression Ratio............................................................27
1.3 Evolving Engine Technologies...................................................................31
1.3.1 Fuels and Compression Ratio ...........................................................................
. . ...........................
.......
1.3 .2 B o ost.............................................................................
1.3.3 Em issions Control ..............................................................................................
1.3.4 Charge and Combustion.....................................................................................
1.3.5 Engine Control............................................
1.3.6 Modern Engine Performance..............................................................................
5
13
.21
21
22
25
31
31
32
33
34
34
1.4 The Plasmatron Engine System........................................................................36
36
1.4.1 PlasmatronDesign and Operation.....................................................................
38
1.4.2 Benefits of Hydrogen and CarbonMonoxide Enhancement..............................
38
1.5 Previous Work ..................................................................
1.5.1 Effects ofAir-Fuel Ratio and Fuel Reformate on Knock andAutoignition........38
1.5.2 Effects of OperatingConditionson Knock and Autoignition............................ 39
39
1.5.3 Effects of Compression Ratio on Efficiency ......................................................
....... 41
1.6 Objectives ..............................................................-.
Chapter 2. Experimental Method ..................................................................
43
2.1 Engine Setup................................................................43
2.1.1 Engine and Dynamometer Specifications...........................................................43
45
2.1.2 A ir andFuel Supply Systems..............................................................................
.... 46
2.2 Engine Control and Measurements ...................................................
46
2.2.1 Engine Control Unit ............................................................................................
47
2.2.2 Fuel Flow Measurement................................................................................
2.2.3 Intake PressureMeasurement and Control.........................................................49
49
2.2.4 Temperature Measurement and Control...........................................................
. 49
2.2.5 A ir Flow Measurement..................................................................................
2.2.6 A ir-FuelRatio Measurement...........................................................................-49
50
2.2.7 Cylinder PressureMeasurement.......................................................................
50
2.2.8 Knock Detection...........................................
2.3 Changing Compression Ratio...........................................................51
7
2.4 Efficiency Experim ents ........................................................................................
51
2.4.1 Experimental Procedure....................................................................................
51
2.4.2 Fuel
........................................................
52
2.4.3 Operating Conditions and CompressionRatio .................................................
52
2.5 Knock Limited Minimum Spark Retard Experiments....................52
2.5.1 ExperimentalProcedure....................................................................................
53
2 .5 .2 Fuels .......................................................................................................................
54
2.5.3 OperatingConditions and Compression Ratio .................................................
56
Chapter 3. Experim ental Results ..............................................................................
59
3.1 Efficiency Results...............................................................................................
59
3.1.1 Effects ofAir-Fuel Ratio andLoad ...................................................................
59
3.1.2 Effects of CompressionRatio .............................................................................
61
3.2 Knock Lim ited Perform ance Results..................................................................
70
3.2.1 Ignition Timing .......................................................................................................
3.2.2 Effects ofA ir-FuelRatio....................................................................................
3.2.3 Effects of Boost .......................................................................................................
3.2.4 Effects of PlasmatronReformate Addition..........................................................
3.2.5 Effects of CompressionRatio .............................................................................
Chapter 4. Chem istry Modeling ..............................................................................
70
71
72
73
75
95
4.1 Background..............................................................................................................95
4.2 M odel Description...............................................................................................
96
4.3 Governing Equations ..........................................................................................
96
4.4 Engine Knock Sim ulations .................................................................................
98
4.4.1 Selection of Input Parameters...........................................................................
4.4.2 Pressure-TemperatureProfiles............................................................................
99
102
4.4.3 Model Sensitivity ..................................................................................................
106
4.4.4 Comparisonto ExperimentalResults...................................................................107
4.5 Constant Volum e Sim ulations...............................................................................112
4.5.1 Effects of Hydrogen Addition ...............................................................................
113
4.5.2 Effects of Carbon Monoxide Addition..................................................................117
Chapter 5. Engine Optim ization ................................................................................
121
5.1 Perform ance Analysis............................................................................................121
5.1.1 CompressionRatio Optimization .........................................................................
121
5.1.2 Fuel Octane Quality and Reformate Addition......................................................123
5.2 Efficiency Analysis ...............................................................................................
126
5.2.1 Effect of CompressionRatio on Brake Efficiency ................................................ 126
5.2.2 Effect of Boost on Brake Efficiency......................................................................128
5.3 Im plications for Engine Design .............................................................................
Chapter 6. Conclusions..............................................................................................
8
130
131
References ..............................................................................
Appendix A
Appendix B
..... ......
...
133
137
M odified Piston Dimensions ..........................................................
Additional Charts for Effects of Air-Fuel Ratio on Knock ............ 141
Appendix D
Additional Charts for Effects of Boost on Knock .......................... 147
Additional Charts for Effect of Fuel Reformate on Knock ............ 155
Appendix E
Additional Charts for Effect of Compression Ratio on Knock ...... 163
Appendix C
9
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10
LIST OF TABLES
Table 1.1 - Outlet composition of an ideal plasmatron for a fuel H/C ratio of 1.9 and of a
37
typical plasm atron [8]............................................................................................
Table 1.2 - Effect of Engine Operating Conditions on Borderline Knock [16]...........40
Table 2.1 - Test Engine Specifications .......................................................................
43
Table 2.2 - Summary of experimental engine transducers and gauges .........................
48
55
Table 2.3 - U TG fuel properties ...................................................................................
Table 4.1 -Temperatures at intake valve close estimated by WAVE engine model.......102
Table 4.2 - Sensitivity of autoignition time to model input parameters. AMign-pp is the
distance in crank angle degrees from autoignition to peak-pressure.......................108
Table 5.1 - Calculated compression ratio for peak torque at 1500 rpm..........................123
Table 5.2 - Effect of downsizing with increased rc on brake efficiency.........................127
Table 5.3 - Effect of boosting and downsizing on mid-load brake efficiency................129
11
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12
LIST OF FIGURES
23
Figure 1.1 - Examples of typical hydrocarbons ............................................................
Figure 1.2 - Pressure measurements from autoignition of PRFs in a rapid compression
machine; fuel-air equivalence ratio: 0.4, initial temperature: 318K, initial pressure:
24
lbar, com pression ratio: 16:1 [3]..........................................................................
Figure 1.3 - Alkane oxidation scheme proposed by Tanaka et al. [3]..........................24
Figure 1.4 - Impact of spark timing on cylinder pressure, 1500 rpm, WOT, X = 1.0.......26
Figure 1.5 - Indicated efficiency for the ideal gas, constant heat capacity, constant
v olum e cy cle ..............................................................................................................
28
Figure 1.6 - Fuel-air results for indicated fuel conversion efficiency as a function of
compression ratio. Fuel: octene; pi = 1 atm, T, = 388 K, xr = 0.05. [1]...............29
Figure 1.7 - Relative fuel conversion efficiency improvement with increasing
compression ratio at wide-open throttle (data from two sources) [1]....................30
Figure 1.8 - Conceptual diagram of maximum torque curves and normal operating range
of naturally aspirated and boosted downsized engines..........................................32
Figure 1.9 - Approximate maximum BMEP (normalized torque) curves for typical
naturally aspirated and boosted engines ..............................................................
Figure 1.10 - Schematic of the plasmatron, courtesy of A. Rabinovich [8] .................
35
36
37
Figure 1.11 - Conceptual plasmatron engine system configuration ............................
Figure 1.12 - Lower octane fuel is supplied to the engine for audible knock when some
energy is derived from H2 and CO (plasmatron reformate). Data simulates 15% and
30% of the gasoline being reformed in the plasmatron fuel reformer [14]. .......... 39
Figure 1.13 - Brake efficiency improvement with compression ratio for engines of
40
several swept volum es [17]. .................................................................................
Figure 2.1 - Base, medium, and high compression ratio pistons (left to right).............44
Figure 2.2 - Illustration of the 13.4:1 piston and cross sections of the combustion chamber
44
at top center piston position...................................................................................
45
Figure 2.3 -Intake air system .......................................................................................
Figure 2.4 - Liquid fuel system .....................................................................................
46
Figure 2.5 - Gaseous fuel system (simulated plasmatron gas supply)...........................46
Figure 2.6 - Engine control and measurement diagram.................................................47
52
Figure 2.7 - Test matrix for efficiency experiments .....................................................
Figure 2.8 - RON and MON of toluene reference fuels as a function of volume fraction nheptane. Compiled by Shell Global Solutions [18] from ASTM vol 05.04, Table 28
56
and internal data......................................................................................................
Figure 2.9 - Test matrix for knock experiments ............................................................
57
13
Figure 3.1 - Change of net indicated efficiency with lambda; 4.Obar NIMEP..............61
Figure 3.2 - Normalized change of net indicated efficiency with lambda; 4.Obar NIMEP.
62
...................................................................................................................................
Figure 3.3 - Change of gross efficiency with lambda; 4.Obar NIMEP..........................62
Figure 3.4 - Combustion event timing; 4.Obar NIMEP, r, = 9.8:1.................................63
Figure 3.5 - Combustion event timing; 4.Obar NIMEP, k = 1.6....................................63
Figure 3.6 - Change of net indicated efficiency with NIMEP; k = 1.0. ........................
64
Figure 3.7 - Change of net indicated efficiency with NIMEP; k
64
=
1.3. ........................
Figure 3.8 - Normalized change of net indicated efficiency with NIMEP; k = 1.0. ......... 65
Figure 3.9 - Normalized change of net indicated efficiency with NIMEP; k = 1.3. ......... 65
Figure 3.10 - Change of gross indicated efficiency with NIMEP; k = 1.0....................66
Figure 3.11 - Change of gross indicated efficiency with NIMEP; k
=
1.3....................66
Figure 3.12 - Change of net indicated efficiency with compression ratio for a range of k;
4.Obar N IME P ........................................................................................................
67
Figure 3.13 - Normalized change of net indicated efficiency with compression ratio for a
range of k; 4.Obar N IM EP. ....................................................................................
67
Figure 3.14 - Change in net indicated efficiency with compression ratio for a range of
loads; k = 1.0. ........................................................................................................
. 68
Figure 3.15 - Change in net indicated efficiency with compression ratio for a range of
load s; k = 1.3 . ........................................................................................................
. 68
Figure 3.16 - Normalized change in net indicated efficiency with compression ratio for a
range of loads; =1 .0 ............................................................................................
69
Figure 3.17 - Normalized change in net indicated efficiency with compression ratio for a
range of loads; k= 1.3 ............................................................................................
69
Figure 3.18 - Change of NIMEP with spark timing for a range of k; rc= 9.8:1, 1500 rpm.
...................................................................................................................................
77
Figure 3.19 - Normalized decrease of NIMEP with spark retard for a range of k; rc =
9.8:1, 1500 rpm, 10.1 bar NIMEPMBT- .------------------.-------................................--------
77
Figure 3.20 - Normalized decrease of NIMEP with spark retard for a range of ) and rc;
toluene fuel, 1500 rpm, MAP at k > 1.0 boosted to match MBT NIMEP at
unboosted WO T X = 1.0. .......................................................................................
78
Figure 3.21 - Change of combustion retard with spark retard for a range of k; rc = 9.8:1,
1500 rpm, 10.1 bar NIMEPMBT
...........------..---------.....-..-- .....--.....- 78
Figure 3.22 - Change of normalized NIMEP with spark retard for a wide range of
operating conditions and compression ratios; toluene fuel except where noted, 1500
rp m .............................................................................................................................
79
14
Figure 3.23 - Change of normalized NIMEP with combustion retard for a wide range of
operating conditions and compression ratios; toluene fuel except where noted, 1500
.....-7 9
rp m ................................................................................................................------.
Figure 3.24 - Averaged heat release rate profiles for several spark timings for toluene and
for TRFs with octane numbers that result in near knocking conditions; rc = 9.8:1, X =
80
1.3, 1500 rpm, MAP = 1.23bar (for NIMEPMBT= 10.1bar). ................................
Figure 3.25 - Change of combustion retard with spark retard for toluene and for TRFs
with octane numbers that result in near knocking conditions; rc = 9.8:1, X = 1.3, 1500
80
rpm, MAP = 1.23bar (for NIMEPMBT= 10.1 bar). ................................................
Figure 3.26 - Effects of X and PRF fuel RON on combustion retard to just avoid knock;
1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted W O T ......................................................................................................
81
Figure 3.27 - Effects of X and TRF fuel RON on combustion retard to just avoid knock;
1500 rpm, MAP at %> 1.0 boosted to match MBT NIMEP at stoichiometric
81
unboosted WO T ....................................................................................................
Figure 3.28 - Effect of X on combustion retard to just avoid knock for UTG fuels; r, =
11.6:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
82
unboosted W OT.............................................................................................
Figure 3.29 - Change of combustion retard to just avoid knock with increasing X; rc =
9.8:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted W OT........................................................................................82
Figure 3.30 - Change of combustion retard to just avoid knock with increasing /; rc =
11.6:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
83
unboosted W O T ......................................................................................................
Figure 3.31 - Change of combustion retard to just avoid knock with increasing X; rc =
13.4:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
.83
unboosted WO T ...................................................................................................
Figure 3.32 - Increase of combustion retard to just avoid knock with increased NIMEP
from boosting for PRFs; rc = 11.6:1, X = 1.0, 1500 rpm........................................84
Figure 3.33 - Increase of combustion retard to just avoid knock with increased NIMEP
from boosting for TRFs; rc = 11.6:1, X = 1.0, 1500 rpm........................................84
Figure 3.34 - Lines of constant MAP calculated from Eq. (3.1) imposed on data from
Figure 3.33. Horizontal lines are drawn at NIMEPMBT. Vertical distance between
curves and lines is torque loss from retarding spark to avoid knock.....................85
Figure 3.35 - Effects of boosted NIMEP and PRF fuel RON on combustion retard to just
85
avoid knock; X = 1.0, 1500 rpm .............................................................................
Figure 3.36 - Effects of boosted NIMEP and TRF fuel RON on combustion retard to just
86
avoid knock; X = 1.0, 1500 rpm .............................................................................
15
Figure 3.37 - Increase of combustion retard to just avoid knock with increased NIMEP
from boosting for UTG96, PRF95 and TRF95; rc = 11.6:1, A = 1.0, 1500 rpm........86
Figure 3.38 - Increase of combustion retard with increased airflow rate from boosting at
1500 rpm, PRFs at k = 1.0 and ) = 1.3, TRFs at k = 1.0 and k = 1.3, and UTG96 at
k = 1.3. Data is from all three compression ratios...............................................
87
Figure 3.39 - Increase of combustion retard with increased NIMEP from boosting at
1500 rpm, PRFs at k = 1.0 and k = 1.3, TRFs at k = 1.0 and k = 1.3, and UTG96 at
k = 1.3. Data is from all three compression ratios...............................................
87
Figure 3.40 - Decrease of combustion retard to just avoid knock and associated NIMEP
increase with increased reformed fuel fraction for PRFs; r, = 11.6:1, k = 1.0, 1500
rpm, MAP for 40% boost (NIMEPMBT 1.4 times unboosted NIMEPMBT). ----......... 88
Figure 3.41 - Decrease of combustion retard to just avoid knock and associated NIMEP
increase with increased reformed fuel fraction for TRFs; rc = 11.6:1, k = 1.0, 1500
rpm, MAP for 40% boost (NIMEPMBT 1.4 times unboosted NIMEPMBT) .----......... 88
Figure 3.42 - Increase of combustion retard with boost and decrease of combustion retard
with reformate fraction for TRF95; r, = 11.6:1, k = 1.0, 1500 rpm. Curve for
NIMEP vs. combustion retard at 40% boost calculated from Eq. (3.1). ............... 89
Figure 3.43 - Increase of combustion retard with rc and decrease of combustion retard
with reformate fraction for TRF95; k = 1.3, 1500 rpm, MAP set to match
NIMEPMBT at unboosted WOT k = 1.0. Curve for NIMEP vs. combustion retard for
89
rc = 13.4:1 calculated from Eq. (3.1)......................................................................
Figure 3.44 - Decrease of combustion retard to just avoid knock with increased reformed
fuel fraction for UTG fuels with PRFs and TRFs for comparison; rc = 11.6:1, k =
1.0, 1500 rpm, MAP for 40% boost (NIMEPMBT 1.4 times unboosted NIMEPMBT).90
Figure 3.45 - Effects of reformed fuel fraction and PRF fuel RON on combustion retard
to just avoid knock; k = 1.0, 1500 rpm, MAP for 40% boost (NIMEPMBT 1.4 times
unboosted NIM EPMBT). ----...........................--------..................................................
90
Figure 3.46 - Effects of reformed fuel fraction and TRF fuel RON on combustion retard
to just avoid knock; k = 1.0, 1500 rpm, MAP for 40% boost (NIMEPMBT 1.4 times
unboosted NIM EPMBT). ..------------------. --------------------..................................................
91
Figure 3.47 - Decrease of combustion retard with increased reformed fuel fraction at
1500 rpm, PRFs at k = 1.0 (40% boost) and k = 1.3, TRFs at k = 1.0 (40% boost)
and %=1.3,and UTGs at k = 1.3. Data is from all three compression ratios........91
Figure 3.48 - Increase of unboosted NIMEP at MBT spark timing with r. for three fuel
types; k = 1.0, 1500 rpm. Closed symbols represent the first high load runs with
new pistons, open symbols represent runs made after several engine hours at high
lo ad . ...........................................................................................................................
92
Figure 3.49 - Normalized average increase of WOT NIMEP at MBT timing with rc; %=
1.0, 1500 rpm. Raw data is from Figure 3.48 .....................................................
92
16
Figure 3.50 - Increase of combustion retard to just avoid knock and change of NIMEP
with increased rc for PRFs; k = 1.0, 1500 rpm, unboosted WOT..........................93
Figure 3.51 - Increase of combustion retard to just avoid knock and change of NIMEP
with increased r, for TRFs; k = 1.0, 1500 rpm, unboosted WOT. ........................ 93
Figure 3.52 - Curves for NIMEP vs. combustion retard from Eq. (3.1) imposed on data
from Figure 3.51. Vertical distance between curves and horizontal lines at
NIMEPMBT is torque loss from retarding spark to avoid knock. .......................... 94
Figure 3.53 - Average increase in combustion retard to avoid knock over all operating
conditions considered, organized by fuel type and compression ratio interval. Bars
94
represent one standard deviation. .........................................................................
Figure 4.1 - Structure of an autoignition simulation. ....................................................
97
Figure 4.2 - A set of 90 cycles of pressure data taken under near knocking conditions
with PRF90; 1500rpm, WOT, X = 1.0. The solid lines are cycles that have a location
of 50% mass fraction burned that is earlier than 90% of the cycles........................100
Figure 4.3 - Residual fractions calculated by a Ricardo WAVE engine simulation; X =
1.3, 1500 rpm , M A P = 1.22 bar...............................................................................101
Figure 4.4 - Comparison of measured cylinder pressure to polytropic compression to find
effective intake valve closing time; 1500 rpm, X = 1.5. .......................................... 101
Figure 4.5 - Experimental pressure profiles and predicted non-reacting end-gas
temperature profiles for stoichiometric and lean air-fuel ratios; iso-octane fuel, 1500
rpm, rc = 9.8:1, NIMEP = 10.1 bar, MBT spark timing. ........................................ 103
Figure 4.6 - Experimental pressure profiles and predicted non-reacting end-gas
temperature profiles for unboosted and boosted conditions; iso-octane fuel, 1500
rpm , rc = 9.8:1, k = 1.0, M BT spark tim ing.............................................................104
Figure 4.7 - Experimental pressure profiles and predicted non-reacting end-gas
temperature profiles for low and high compression ratios; iso-octane fuel, 1500 rpm,
104
M A P = 1 bar, M BT spark tim ing. ...........................................................................
Figure 4.8 -Predicted non-reacting end-gas temperature profiles with the same pressure
profile for PRF100 and TRF100 fuels; 1500 rpm, rc = 9.8:1, MAP = 1.0 bar, k = 1.0,
M B T spark tim ing....................................................................................................105
Figure 4.9 - Predicted end-gas temperature profiles with the same pressure profile for
simulations with reaction kinetics disabled and enabled for PRF95; 1500 rpm, r, =
9.8:1, M AP = lbar, X = 1.0, Osp = 60 BTC...............................................................106
Figure 4.10 - Effect of initial temperature on predicted autoignition time in increments of
5 K for PRF95; 1500 rpm rc = 9.8:1, MAP = Ibar, X = 1.0, Osp = 6' BTC. ............ 107
Figure 4.11 - Pressure traces from several spark timings and corresponding simulated
end-gas temperature profiles; 1500 rpm, rc = 11.6:1, k = 1.0, MAP = 1.Obar. 4 BTC
is interpreted as the predicted knock limited spark timing......................................109
17
Figure 4.12 - Comparison of model predicted combustion retard to experimental
combustion retard with increasing compression ratio for PRF95; 1500 rpm, X = 1.0,
109
M AP =1 .Obar. .........................................................................................................
Figure 4.13 - Comparison of model predicted combustion retard to experimental
combustion retard with increasing compression ratio for TRF95; 1500 rpm, X = 1.0,
1 10
M AP = 1.0 b ar. ........................................................................................................
Figure 4.14 - Comparison of model predicted combustion retard to experimental
combustion retard with increasing air-fuel ratio for PRF95; 1500 rpm, rc = 9.8:1,
N IM E PM BT
10.1 bar. .............................................................................................
110
Figure 4.15 - Comparison of model predicted combustion retard to experimental
combustion retard with increasing air-fuel ratio for TRF95; 1500 rpm, rc = 9.8:1,
N IM E P M BT = 10.1 bar..............................................................................................111
Figure 4.16 - Comparison of model predicted combustion retard to experimental
combustion retard with increasing boosted NIMEP for PRF95; 1500 rpm, r, = 9.8:1,
11 1
X = 1 .0 ......................................................................................................................
Figure 4.17 - Comparison of model predicted combustion retard to experimental
combustion retard with increaing boosted NIMEP, PRF95; 1500 rpm, rc = 9.8:1, X =
1 12
1 .0 . ...........................................................................................................................
Figure 4.18 - Temperature and OH concentration for constant-volume autoignition
simulations of iso-octane and of iso-octane with 5.5% H 2 by energy; Tinit = 875 K,
P init = 4 5 bar, k = 1.5................................................................................................113
Figure 4.19 - Temperature and OH concentration for constant-volume autoignition
simulations of iso-octane with 5.5% H2 by energy, and of iso-octane with simulated
non-reactive H 2 (nrH 2); Tinit = 875 K, Pinit = 45 bar, X = 1.5....................................114
Figure 4.20 - OH consumption by reactions with fuel molecules for constant-volume
autoignition simulations of iso-octane with 5.5% H2 by energy, and of iso-octane
with simulated non-reactive H 2 (nrH 2); Tinit = 875 K, P init= 45 bar, k = 1.5. .......... 115
Figure 4.21 - Rates of reaction of OH and H with iso-octane molecules for constantvolume autoignition simulations of iso-octane with 5.5% H 2 by energy, and of isooctane with simulated non-reactive H2 (nrH 2); Tinit = 875 K, Pinit= 45 bar, k = 1.5.
.................................................................................................................................
115
Figure 4.22 - Temperature and OH concentration for constant-volume autoignition
simulations of iso-octane with 5.5% H2 by energy, and of iso-octane with simulated
H 2 (prH2 ) that reacts only with OH; Tinit = 875 K, Pinit = 45 bar, X = 1.5.................116
Figure 4.23 - Proposed mechanism by which hydrogen impacts the autoignition of alkane
hydrocarbons. Base diagram is from Tanaka et al. [3]...........................................117
Figure 4.24 - Comparison of the reaction rate constant for CO+02 => CO2+O from three
sources. The rate from Scire et al. [26] was used in this study...............................118
18
Figure 4.25 - Temperature and OH concentration for constant-volume autoignition
simulations of iso-octane and of iso-octane with 7% CO by energy; Tiit = 875 K,
P init= 45 bar, k = 1.5................................................................................................118
Figure 4.26 - Temperature and OH concentration for constant-volume autoignition
simulations of iso-octane with 7% CO by energy, and of iso-octane with simulated
non-reactive CO (nrCO); Tinit = 875 K, Pinit = 45 bar, k = 1.5. ................................ 119
Figure 5.1 - Tradeoff between NIMEP and compression ratio. Values shown
approxim ate TRF95 fuel..........................................................................................122
Figure 5.2 - Contours of constant combustion retard for varying reformate fraction and
fuel RON; k = 1.0, 1500 rpm, MAP for 40% boost. Data from Figure 3.45 (PRFs,
124
left) and Figure 3.46 (TRFs, right). .........................................................................
Figure 5.3 - Contours of constant combustion retard for varying boosted NIMEP and fuel
RON; k = 1.0, 1500 rpm. Data from Figure 3.35 (PRFs, left) and Figure 3.36
12 5
(TRF s, right). ...........................................................................................................
Figure 5.4 - Contours of constant combustion retard for varying compression ratio and
fuel RON; k = 1.0, 1500 rpm, MAP = 1.0 bar.........................................................125
Figure 5.5 - Estimated increase of mid-load brake efficiency with re, with and without
downsizing the engine to maintain constant maximum torque output; 1500 rpm, k =
1.0, 2.6 bar baseline B M EP.....................................................................................128
Figure 5.6 - Estimated increase of mid-load net and brake efficiencies with boosting and
downsizing; 1500 rpm, k = 1.0, 2.6 bar baseline BMEP.........................................129
19
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20
CHAPTER 1. INTRODUCTION
Knock in spark ignited internal combustion engines is the external sound produced by
pressure oscillations in the cylinder caused by spontaneous autoignition of the air-fuel
mixture before it is consumed by the propagating flame front [1]. The knocking noise is
unacceptable in today's refined vehicles, and repeated exposure to the extreme local
pressures and temperatures caused by very rapid autoignition are damaging to engine
components.
Since the inception of the spark ignited internal combustion engine, knock has been one
of the fundamental limiting factors. Historically, its main effect has been to limit the
maximum compression ratio, and thus the efficiency, of automotive engines. As fuel
quality and engine technology have improved, engine compression ratios have increased
to the point that further increases result in smaller efficiency improvements. Currently,
the main effect of knock is reduced low speed torque - due to spark retard on naturally
aspirated engines, and limited boost on turbocharged or supercharged engines. The
motivation of this work is to investigate the tradeoffs between fuel type, compression
ratio, boost level, air-fuel ratio, and spark timing at the knock limit such that informed
decisions can be made during the selection of new engine configurations.
1.1 KNOCK IN SPARK IGNITION ENGINES
The term "knock" describes the sound resulting from the spontaneous ignition of the
unburned air-fuel mixture ahead of the advancing turbulent flame front. The unburned
mixture is compressed by the expansion of the burned gases produced in the turbulent
flame propagating from the spark plug outwards. The increases in unburned mixture
temperature and density associated with this pressure rise initiate the chemical reactions
that lead to autoignition. If the flame does not consume the unburned fuel, or "end-gas",
before the chemical reactions proceed to the point of rapid bulk heat release, the resulting
autoignition causes local temperatures and pressures to increase sharply. Simulations
performed at Leeds University [2] indicate that the local pressure may rise to upwards of
300 atmospheres in as little as one microsecond. A pressure wave then propagates
outwards from the location of autoignition at the local speed of sound and resonates
within the engine cylinder. The pressure oscillations cause vibration of the head and
block, resulting in a clanging or pinging sound. While light autoignition can be
21
functionally acceptable, audible knock is disagreeable, and the high local temperatures
and pressures associated with heavy knock cause engine damage.
Since knock can damage the engine and produces a disagreeable sound, engine design
and operating parameters are adjusted for its avoidance. Knock is mainly a problem
during high load, which creates high peak pressures, and low engine speed, which
provides a long residence time for the end-gas. To avoid knock under these conditions
engine control systems typically retard spark timing to lowers peak pressure, but high
spark retard results in a significant decrease in engine torque and combustion stability
and an increase in exhaust temperatures. To avoid the need for excessive spark retard,
compression ratio is limited, and for turbocharged or supercharged engines boosted inlet
pressure is limited.
1.1.1 Autoignition Chemistry
Gasoline consists of approximately 60% alkane hydrocarbons and 30% aromatic (cyclic)
hydrocarbons, with the remainder composed of alkenes, alcohols, and additives. Figure
1.1 shows some examples of typical hydrocarbons. The low temperature (less than
1000K) chemistry of alkanes larger than C5 are most relevant to gasoline autoignition, so
they are the topic of most work on the subject. Figure 1.2 shows the evolution of
pressure in the combustion chamber of a rapid compression machine for several mixtures
of primary reference fuels (iso-octane and n-heptane). Two-stage autoignition, an
important characteristic of alkane autoignition, is clearly visible. The increasing duration
of the second stage with octane number results in an overall increase in autoignition time.
Three sequences are responsible for the alkane autoignition process.[3] A highly
exothermic chain-branching reaction sequence is responsible for the first stage:
RH+OH+202
=>P
+20H
In this sequence of reactions, a hydroxy radical combines with oxygen and a fuel
molecule to produce two hydroxy radicals, which carry on the sequence. As the
temperature rises, a weakly exothermic competing reaction sequence slows down the
temperature rise, causing a transition to the second stage:
2RH+ 20H + Q2 = P +H202
22
In this sequence of reactions hydroxy radicals are tied up into hydrogen peroxide.
Finally, as temperature and radical concentrations rise, decomposition of the hydrogen
peroxide initiates rapid exothermic combustion (autoignition):
H2 0 2 ->2OH
Alkanes:
CH3 C
I
I
CH3 -C-CH 2 C-CH 3
C
CH3
CH3-CH 2-CH2 -CH 2-CH 2-CH 2 -CH 3
Isooctane
(2,2,4-Trimethylpentane)
ON=100
Aromatics:
n-Heptane
ON=0
/CH 2
/CH 3
/CH-C\
H
CH
CH-CH
Toluene
RON-120
/CH-C
CH
H
CH 3
CH-CH
Ethylbenzene
Figure 1.1 - Examples of typical hydrocarbons
Figure 1.3 shows the three sequences in an overall autoignition scheme. The lower loop
is the chain branching sequence. The fuel molecule is attacked by a hydroxy radical and
then combines with oxygen and undergoes internal isomerization before releasing two
hydroxy radicals. The upper loop is the high temperature mechanism. At higher
temperatures, after being attacked by the hydroxy radical, the alkyl radical reacts with
oxygen to produce olefin and an HO 2 radical, which combines to produce hydrogen
peroxide. As temperatures rise, the built up hydrogen peroxide molecules dissociate into
hydroxy radicals and autoignition occurs.
Although the typical trend is for decreasing time to autoignition with increased
temperature, variations in the rates of intermediate reactions result in complex variations
of autoignition time. For alkanes such as n-heptane and iso-octane there is a temperature
interval (roughly 750 K to 850 K), called the Negative Temperature Coefficient (NTC)
region, for which autoignition time increases with temperature. Mixtures of fuels show
more complicated variations of autoignition time with temperature. Although
23
experimental ignition times give an indication of a fuel's autoignition properties, due to
transient pressures and temperatures in SI engine end-gas, they do not relate directly to
knock propensity.
Prm---e-r---F-ldTac
B
(K)
7
1393
6
1194
0~
4
31
5~.995
75
90100
7so-Ocane)
ON=O0
(n-Heptane
795
597
2
398
1
19
0
.5
5
10
time (msec)
Is
2
Figure 1.2 - Pressure measurements from autoignition of PRFs in a rapid
compression machine; fuel-air equivalence ratio: 0.4, initial temperature: 318K,
initial pressure: lbar, compression ratio: 16:1 [3].
High Temperature
-H02
H202
Olefin
02
----- - -
R-----RH+0
2R
V-
M
RH
-
H O
H20
02
IROO
hr erazakat
Low
Temperature
-ROOH
OH
02
-OOROOH
OROOH
*ORO
Figure 1.3 - Alkane oxidation scheme proposed by Tanaka et al. [3]
24
The system most commonly used to rate the knock propensity of SI engine fuels is the
octane rating method. Fuels are compared to Primary Reference Fuels (PRFs) in a
standardized single cylinder engine under two operating conditions. American Society
for Testing and Materials standard ASTM D-2699 specifies the research method, which
measures the RON (Research Octane Number). ASTM D-2700 specifies the motoring
method, which gives the MON (Motoring Octane Number). For each test the
compression ratio is increased until the engine just knocks with the fuel under test. The
Octane Number (ON) of the fuel is the ON of the PRF that just knocks under the same
conditions. The motoring method is performed with higher inlet temperature, higher
speed, and more spark advance than the research method. For most fuels, knock
propensity increases more with severity of operating conditions than it does for PRFs,
thus the MON is typically lower than the RON.
1.1.2 Factors Affecting Autoignition in SI Engines
Under most conditions heat release from the end-gas prior to autoignition is small, so
compression and normal combustion dictate the evolution of cylinder pressure. Cylinder
pressure, however, has a significant effect on end-gas thermodynamic state and reaction
rates. An estimate of how pressure affects end-gas temperature can be gained by looking
at the P-T relationship for adiabatic compression:
T=
Tinitial
(
(
Pinitial /
Higher pressures result in higher temperatures and increased reaction rates. Factors that
affect cylinder pressure are:
Inlet pressure - Controls the amount of air-fuel mixture entering the cylinder. Higher
inlet pressure increases engine torque output and increases cylinder pressures.
A ir-fuel ratio- Affects combustion rates and the energy released during combustion.
Dilution from excess air and changes in thermodynamic properties of the mixture with
air-fuel ratio affect how the heat release impacts pressure evolution.
Spark timing - Affects combustion phasing. Late combustion phasing results in lower
cylinder pressures because the bulk of the combustion process occurs in a larger, faster
expanding volume. Figure 1.4 shows how cylinder pressures change with spark timing.
25
60
Spark Timing
50-
50% mass...
40 -
burned
30
-
-16deg
-8deg
BTC
-4deg
ATC
BTC (MBT)
-
20
10
-
Opark
A
-45
0
Crank Angle
45
90
(0ATC)
Figure 1.4 - Impact of spark timing on cylinder pressure, 1500 rpm, WOT, X = 1.0
Compression ratio- The ratio of the maximum cylinder volume to minimum cylinder
volume. As compression ratio is increased the cylinder volume before and during
combustion becomes smaller, resulting in higher cylinder pressures.
Engine speed - The duration of the engine cycle scales inversely with engine speed. As
speed increases, the duration of time for which unburned end-gas is subjected to high
pressures and temperatures decreases.
Chargepreparation- Turbulence and mixture homogeneity affect speed of combustion.
As combustion duration decreases the end-gas is consumed faster, resulting in less time
for autoignition reactions to occur.
Combustion chamber geometry -Longer flame travel distances result in longer burn
duration and more time for autoignition to occur.
Factors that affect the evolution of temperature and the chemical composition of the endgas in response to the evolution of pressure are:
26
Mixture composition - The ratio of specific heats (y) is an important parameter in Eq. (
1.1). Higher y results in higher compression temperatures. Air has a higher ratio of
specific heats (y = 1.4) than gasoline (y~1.0) and burned gas (y~1.3), so increased air-fuel
ratio increases y and increased residual fraction reduces y. Reactant concentrations, fuel
types, and fuel additives also have important effects on autoignition reaction rates.
Initial temperature- End-gas compression temperature scales closely with initial
temperature, as indicated by Eq. ( 1.1). Higher initial temperatures result in higher
unburned gas temperatures throughout the compression and combustion processes,
increasing the rates of autoignition reactions. Initial temperature is affected by inlet air
temperature, heat transfer, and the temperature and quantity of residual gas.
Heat transfer during compression and combustion affects both the evolution of pressure
and the temperature of the end-gas. During the first part of compression heat is
transferred to the unburned gas. During the second part of compression and during
combustion heat transfer from the burned gas reduces cylinder pressure and heat transfer
from the end-gas reduces its temperature and slows autoignition reactions
1.2 EFFICIENCY EFFECTS OF COMPRESSION RATIO
To avoid knock, the maximum compression ratio of an engine must be limited. The
improvement of engine efficiency with compression ratio is a key motivation behind
work to improve knock performance. This section gives an overview of the effect of
compression ratio on engine efficiency.
An illustrative model of an SI engine cycle is the ideal gas, constant heat capacity,
constant volume heat addition cycle. In this cycle the working fluid is compressed
isentropically, heat is added at constant volume, and then work is extracted as the
working fluid is isentropically expanded. A constant volume heat rejection (fluid
exchange) process completes the cycle. For this cycle the efficiency is:
qf/ig
=
1-
1
(1.2)
rrl
The results of this equation are shown in Figure 1.5. Efficiency increases with 7 and
increases with rc with decreasing returns at high rc. This cycle is illustrative, but it is
missing many real-engine effects.
27
0.75
y = 1.4
0.7 -
(air)
0.65 y = 1.3
0.6 0.55 '
0.5 y=1.2
0.45 0.4 0.35 0.3 0.25
0
5
10
15
20
25
30
rc
Figure 1.5 - Indicated efficiency for the ideal gas, constant heat capacity, constant
volume cycle
For a more accurate estimate of real engine performance, the effects of fluid properties
and combustion chemistry must be taken into account. The working fluid does not have
constant specific heats and its composition changes during the cycle. Combustion does
not go to completion - the burned gases are approximately in equilibrium at high
temperatures and have a frozen composition at low temperatures. The fuel-air cycle takes
these effects into account [1].
The results of the air-fuel cycle can be calculated using fits to thermodynamic data and
equilibrium computer code. Figure 1.6 shows indicated efficiency as a function of
compression ratio for a series of fuel/air equivalence ratios ( = (F/A)/(F/A)stoich). Lines
of constant j are slightly lower than those predicted by Eq. ( 1.2) from the ideal cycle.
Efficiency increases with decreasing
4 due to an increased value of burned gas
y (from
decreased temperature and an increased proportion of diatomic gases). The higher y
causes the burned gases to expand through a greater temperature ratio for a given
expansion ratio, increasing work output [1]. Efficiency decreases with increasing * on the
rich side because the fuel is not fully oxidized due to insufficient air.
28
0.65
--
0.4
0.5
1.0 at0.6
-.
- =388K
p, =
0.60
81.0
x, = 0.05
-
-
--
0. 55 -
0.50
1.2
-
7)f i 0.45 ----
40
-
0. 35
-
0.25
1.4
//
12'1
0
5
10
I
20
15
Compression ratio r
25
30
Figure 1.6 - Fuel-air results for indicated fuel conversion efficiency as a function of
compression ratio. Fuel: octene; p1 = 1 atm, T1 = 388 K, xr = 0.05. [1]
In a real spark ignited engine several other factors impact the effect of compression ratio
on efficiency. Figure 1.7 shows a comparison of real engine performance (circa 1960,
there is a lack of data for a wide range of rc in modem engines) and the fuel-air cycle
results. At low rc the experiments match the air-fuel cycle well. At high rc, heat transfer,
crevice effects, and friction cause efficiency to decrease.
Under normal operation, heat transfer from the combustion chamber corresponds to
roughly 30% of the fuel energy, similar to the fuel conversion efficiency of the engine.
Thus, small changes in heat transfer can have a great impact on engine performance.
Increasing the compression ratio decreases temperatures late in the expansion stroke
because more energy is removed as work. However, increasing rc increases temperatures
during combustion and early expansion, and increases the heat transfer coefficient due to
increases in density and mixture motion. Changes in combustion chamber surface area
also affect heat transfer, but are dependent on engine geometry. The combined effect is
that heat transfer decreases up to a rc of about 10:1 and increases above that [1].
29
1.3
Fuel-air cycle
1.2
Git
93 ;(CN)
1.1-
Wide-open throttle
1.0I.
8
I
10
I
12
I
14
I
16
I
18
I
20
I
22
24
Compression ratio rc
Figure 1.7 - Relative fuel conversion efficiency improvement with increasing
compression ratio at wide-open throttle (data from two sources) [1].
Crevices between the piston and the bore, above and behind the top compression ring,
and around the valves and spark plug comprise approximately 2-3% of the combustion
chamber clearance volume. The fraction of the charge that gets trapped in the crevices
during combustion is even greater because the cold unburned gas in the crevices is much
more dense than the hot burned gas in the cylinder. The flame does not propagate into
the crevices and thus detracts from the energy extracted from the charge during the main
combustion event. As the combustion chamber volume is decreased to increase
compression ratio the crevice volume becomes a greater fraction of the total clearance
volume, and higher cylinder pressures increase the density of the mixture in the crevices.
The higher mass trapped in the crevices leads to reductions in efficiency.
Increasing the compression ratio increases pressures in the cylinder during compression,
combustion, and expansion. This increases the force between the piston and the bore,
which increases sliding friction, and between the piston and the connecting rod, which
increases friction at the wrist pin and crankshaft bearings.
30
1.3 EVOLVING ENGINE TECHNOLOGIES
The spark ignition engine is being steadily improved upon. From 1984 to 2000,
maximum torque normalized by engine displacement has increased by about 1.5% per
year [4]. Engine performance has also been impacted by government emissions
regulations and, more recently, by vehicle fuel efficiency standards. The technologies
employed to increase torque output and meet government regulations has important
implications on engine configuration, operating parameters, and knock control.
1.3.1 Fuels and Compression Ratio
From the 1920s to the 1960s steady improvements in fuel octane quality allowed SI
engine compression ratios to rise steadily [5]. The improvements have been made
through better hydrocarbon selection and processing, and through addition of octane
enhancers such as TEL (tetraethyl lead) and, more recently, oxygenates like MTBE
(methyl tert-butyl ether). Since the 1960s fuel octane quality has been relatively constant
and variations in r, have been in response to changes in regulations and engine
technology. Compression ratios of modem naturally aspirated SI engines are steadily
increasing at about 0.1 units every three years, and are currently between 9.5:1 and 11:1.
1.3.2 Boost
Historically inlet pressure boosting has been used to increase the power output of high
performance engines. Increasingly, most notably in Western Europe, smaller boosted
engines are being used as replacements for larger naturally aspirated engines. The most
common method of boosting inlet pressure is through turbocharging, where energy in the
exhaust gas is extracted by a turbine and used to power an inlet air compressor. Most
modem turbochargers have variable geometry nozzles or turbines for a greater dynamic
range and faster throttle response. An intercooler is usually used with boosted engines to
reduce the temperature of the compressed inlet air before it enters the engine cylinder.
Downsizing an engine requires it to operate at higher specific loads under normal driving
conditions. Figure 1.8 shows the increased BMEP range for a downsized engine (BMEP
is Brake Mean Effective Pressure, a measure of torque output normalized by engine
displacement). Engine downsizing improves efficiency by reducing throttling losses,
friction, and heat transfer. To maintain acceptable maximum torque in a downsized
engine, inlet boosting is required. To avoid knock, however, boosted engines typically
31
require a reduced compression ratio. In practice, boosted downsized engines have
moderate efficiency benefits. If an engine could be boosted without reducing part-load
compression ratio, through methods such as compression variation or fuel reforming, the
efficiency benefits of boosting and downsizing could be improved significantly.
16
CU
~12
E
Normal operating range
for boosted and
downsized engine.
8a)
Normal
operating
range for
0
-Baseline
.4-apiated
engine.
-
NA
Boosted
0
0
3500
7000
Speed (RPM)
Figure 1.8 - Conceptual diagram of maximum torque curves and normal operating
range of naturally aspirated and boosted downsized engines.
1.3.3 Emissions Control
Since the 1960s emissions regulations have imposed important restrictions on SI engine
operation. To meet current emission standards engine designers employ a combination of
emissions reduction techniques. Improvements in air-fuel ratio control and reductions of
crevice volume reduce emissions of hydrocarbons and carbon monoxide. Exhaust Gas
Recirculation (EGR) reduces the temperature of the burned gas and slows NOx
production. Additionally, virtually every SI engine powered vehicle in developed
countries comes with a 99% efficient Three-Way Catalyst (TWC) to oxidize unburned
hydrocarbons (HC) and carbon monoxide (CO) and reduce oxides of nitrogen (NOx).
Efficient operation of a TWC requires restrictions to be placed on engine operating
parameters, the most important of which is a stoichiometric air-fuel ratio. Significant
32
efficiency benefits can be realized by running an engine lean of stoichiometric, but under
oxygen-rich conditions the TWC is unable to reduce NOx. If the engine is run
sufficiently lean, NOx emissions drop to the point that further reductions are not
necessary. With current technology, however, the high air-fuel ratio required is in excess
of the "lean limit" - the limit beyond which combustion stability becomes unacceptable.
Research is underway at MIT [6] and elsewhere to extend the lean limit using charge
motion and fuel reforming. Additionally, lean NOx reduction technology currently under
development for diesel engines may eventually make its way to SI engine applications.
TWCs also place limits on maximum exhaust temperature during full load operation,
limiting maximum spark retard. Enriching the air-fuel ratio at high loads is commonly
used to increase the heat capacity of the burned gas and reduce the exhaust temperature.
TWCs also place restrictions on fuel composition. Impurities such as sulfur and additives
such as octane-enhancing TEL poison catalytic converters and reduce their efficiency
over time.
1.3.4 Charge and Combustion
As described in the previous section, most SI engines run a stoichiometric air-fuel ratio
over much of the engine map. Emissions requirements prevent lean operation, but
enrichment is sometimes used at high loads to decrease knock tendencies and reduce
exhaust temperature. EGR is commonly used for emissions and efficiency benefits.
There are several ways to introduce fuel into the inlet air. The most common is
sequential port injection. In this configuration fuel injectors at the inlet ports are directed
at the back of the intake valves. Increasingly common is Gasoline Direct Injection
(GDI). With GDI, the fuel is injected directly into the cylinder during the intake or
compression strokes. In the USA GDI strategies are limited to stoichiometric, usually
early injection. In places where NOx regulations are not as stringent, lean stratified
charge late injection strategies may be used. Under high loads, early injection while the
intake valve is still open cools the charge. This increases its density, allowing more airfuel mixture into the cylinder, and slows autoignition reactions, allowing a higher
compression ratio or increased boost.
Most modern SI engines employ some kind of charge motion control to improve mixing
and combustion and reduce cycle-to-cycle variability under low load conditions. Charge
motion has also been used to improve knock performance by speeding up combustion,
33
thus reducing end-gas residence time [7], and can be used to extend the lean limit [6].
Methods for introducing charge motion include adjustable vanes in the inlet manifold
close to the intake port, valve masking, and variable valve actuation.
Intake tuning adds significantly to modem SI engine performance. Most engines have
intake runners and plenum volumes tuned to take advantage of pressure waves to increase
the pressure at the valve at the end of the intake stroke. Many engines have variable
intake geometry such that they can be tuned to several engine speeds. Although not as
effective, exhaust tuning also helps to improve volumetric efficiency.
Modem combustion chambers have been optimized for fast combustion. Central spark
plug location and compact combustion chamber geometry reduces the distance that the
flame has to travel and increases flame area. Reduced combustion time results in less
cycle-to-cycle variability under low loads and reduced time for autoignition reactions to
occur at high loads.
1.3.5 Engine Control
The ECU (Engine Control Unit) of a modem engine receives inputs from many sensors
and controls many operating parameters and peripheral systems. The control parameters
most important to engine operation are fuel injection quantity and spark timing. Under
normal part-load operation the ECU executes closed-loop control of the fuel quantity
using feedback from the exhaust 02 sensor. Spark timing is retarded about 5' CA from
MBT to reduce NOx emissions with a very small efficiency penalty. At high loads spark
retard is used to lower cylinder pressures for knock avoidance. The amount of spark
retard is usually based on operating conditions. In some vehicles a knock detection
sensor is employed.
1.3.6 Modern Engine Performance
For a typical spark ignition engine, maximum efficiency occurs at about mid speed,
three-quarters load. The reasons for decreasing efficiency around this point are:
" Decreased speed - at lower speeds there is more time for heat transfer in each cycle.
" Increased speed - at high speeds rubbing and fluid friction become more important.
" Decreased torque - throttling increases pumping work, and the relative importance of
friction increases
34
*
Increased torque - spark retard is required to avoid knock and enrichment is used to
protect the catalytic converter.
Figure 1.9 shows approximations of BMEP as a function of engine speed for typical
naturally aspirated and boosted engines. Maximum torque reaches a peak at mid speed
and drops off in either direction. For naturally aspirated engines the peak is governed by
the amount of air that flows into the engine, which depends on tuning and flow
restrictions. For boosted engines the peak is determined by the size of the turbocharger
and the maximum cylinder pressures that the engine can structurally endure. At lower
speeds the maximum torque drops off due to the spark retard required to avoid knock,
and due to increased heat transfer and decreased intake manifold tuning effects. At
higher speeds the maximum torque is reduced by increased friction and flow restriction.
20
maximum cylinder pressure
(mechanical integrity) and/or
turbocharger performance
IL
E
16
-
airflow restriction
and/or turbocharger
performance
spark retard and
reduced boost (to
avoid knock)
12 limited airflow
through valves and
0
8 -
spark retard (to
avoid knock) and
heat transfer
intake
air flow restriction
through ports and
valves
4-
Boosted
-Baseline
NA
0
0
2000
4000
6000
8000
Speed (RPM)
Figure 1.9 - Approximate maximum BMEP (normalized torque) curves for typical
naturally aspirated and boosted engines
35
1.4 THE PLASMATRON ENGINE SYSTEM
"Plasmatron" is the name used to describe a partial oxidation thermal fuel reformer under
development by the MIT Plasma Science and Fusion Center in cooperation with Arvin
Meritor. The hydrogen rich reformate produced by the plasmatron is beneficial to several
aspects of engine operation.
1.4.1 Plasmatron Design and Operation
The plasmatron, shown schematically in Figure 1.10, uses a plasma arc to partially
oxidize a stream of rich air-fuel mixture. Ideally, all of the carbon in the fuel would be
converted to CO and all of the hydrogen would be converted to H2 :
CmHn + M(O2 + 3.773N 2 )=> mCO+L! H 2 + 3.773-mN
2
2
2
2
(1.3)
In practical applications the partial oxidation is not ideal, resulting in a mixture of several
species in the outlet gases. For a typical plasmatron, about 20% of the fuel heating value
is released as heat and the rest is converted to the chemical energy in CO and H2 . Table
1.1 shows ideal and typical plasmatron outlet compositions.
Fuel
Air I
1-Plas matron
2- 1 Stage Reactor
3-Nozzle Section
4- 2 nd Stage Reactor
Air 2
Air 3
Fuel
Figure 1.10 - Schematic of the plasmatron, courtesy of A. Rabinovich [8]
36
Table 1.1 - Outlet composition of an ideal plasmatron for a fuel H/C ratio of 1.9 and
of a typical plasmatron [81
H2
CO
N2
Ideal Plasmatron
25%
26%
49%
Typical Plasmatron
20%
22%
51%
CO 2 and H2 0
smaller hydrocarbons
0%
0%
6%
~1%
Species
The conceptual configuration of the plasmatron engine system is shown in Figure 1.11.
Gasoline reformate would be used to enhance the air-fuel mixture entering the engine
with H2 and CO. The proportion of fuel delivered to the plasmatron would be modulated
to achieve the required combustion stability or knock avoidance while minimizing the
efficiency penalty. Since the size of the plasmatron is limited by cost and space
constraints, the maximum fuel reformed fraction would likely be limited to 20-30%.
,,,1 "I
3
r
V
air
V
H2 ,CO, gasoline
N 2 (eg.85%)
exhaust
Figure 1.11 - Conceptual plasmatron engine system configuration
37
1.4.2 Benefits of Hydrogen and Carbon Monoxide Enhancement
Hydrogen has two important properties that are beneficial to SI engine performance when
used in conjunction with gasoline:
1)
The laminar flame speed of hydrogen is about three times that of gasoline (at 1000
C, 1 atm, stoichiometric SL,H2= 170 cm/s [9] and SL,gasoine= 45.3 cm/s [1]). This
stabilizes combustion and extends the dilution limit, either with air or with
recirculated exhaust gas [10]. Faster combustion also consumes the unburned
end-gas faster and helps to prevent knock at high loads.
2) Hydrogen has a high octane number. Since it is more knock-resistant than the
highest ON PRF, it cannot be rated on the octane scale as specified by the ASTM
standards. A previous literature review estimated the relative octane number of
hydrogen to be between 130 and 140 [11][12][13]. A rating of 140 showed good
correlation to experimental results when used in a bond-weighted mixing octane
number estimation method [14]. This is partially due to its high flame speed, but
mainly due to its resistance to autoignition. Recent work by Topinka et al. [14]
has shown that replacing 10% of the fuel with H2 (by energy) lowers the required
octane number of the primary fuel by 10 points for the same indicated torque.
Carbon monoxide has a mild effect on increasing laminar flame speed. It is also resistant
to knock and recent tests have indicated an ON of about 106. Experiments by Topinka et
al. [14] have shown that CO improves knock resistance of the primary fuel about half as
much as hydrogen for the same fuel fraction (by energy).
1.5 PREVIOUS WORK
This section summarizes the results of a preceding MIT investigation of the effects of airfuel ratio and fuel reformate on knock. Relevant studies of the effects of operating
conditions on knock, and the efficiency effects of compression ratio are also described.
1.5.1 Effects of Air-Fuel Ratio and Fuel Reformate on Knock and Autoignition
In a preceding study by Topinka et al. [14] [8], the effects of air-fuel ratio and fuel
reformate on the octane number of the PRF required to just avoid knock were
investigated. It was found, contrary to previous studies [15], that at constant indicated
torque the octane requirement of the engine increased slightly with air-fuel ratio (about 2
38
ON for relative air-fuel ratio from 1.1 to 1.7). Through extensive experiments with ideal
fuel reformate addition (25% H2 , 26% CO, balance N2 ), it was found that the octane
number of the PRF main fuel could be reduced as fuel reformate was added. The
magnitude of this reduction is shown in Figure 1.12. It was also found that by adding just
the H2 from the fuel reformate about 50% of the reformate benefit could be achieved, and
by adding just the CO from the fuel reformate about 30% of the reformate benefit could
be achieved. This study also proposed a method for modeling autoignition in SI engines
using a zero-dimensional end-gas simulation with chemical kinetics.
30
-4-high load, lambda=1.1
-*-- high load, lambda=1.3
25
U-
20
-A- high load, lambda=1.5
-- high load, lambda=1.7
-8-mid load, lambda=1.1
15
--E-mid load, lambda=1.3
-A-mid load, lambda=1.5
0
0
(D
0hCU
M
10
0)
5
0
0
0.05
0.1
0.15
0.2
0.25
0.3
Fuel Energy From H2+CO [%/100]
Figure 1.12 - Lower octane fuel is supplied to the engine for audible knock when
some energy is derived from H2 and CO (plasmatron reformate). Data simulates
15% and 30% of the gasoline being reformed in the plasmatron fuel reformer [141.
1.5.2 Effects of Operating Conditions on Knock and Autoignition
Experiments by Russ [16] on a single cylinder engine have been used to estimate the
magnitude of the effects that engine operating conditions and compression ratio have on
the fuel ON required to avoid knock. The results are summarized in Table 1.2.
1.5.3 Effects of Compression Ratio on Efficiency
Although there are several studies on the effects of small compression ratio changes,
there are few on the effects of wide rc sweeps on a modem engine. Muranaka et al. [17]
39
compiled results of r, sweeps from earlier references and added their own data (Figure
1.13). They found that efficiency improvement with rc depended on operating conditions
and swept volume. For a 500 cc cylinder at part load, efficiency initially increases at
about 3% per unit r, at 9:1 and peaks between 13:1 and 15:1. Their analysis indicates
that the major factors for the efficiency limit are heat transfer and unburned fuel.
Table 1.2 - Effect of Engine Operating Conditions on Borderline Knock [16]
Effect on ON requirement
Operating condition
Spark advance
1 ON/iV spark advance
Intake air temperature
1 ON/7 K
Air/fuel ratio
Intake pressure
peaks 5% rich of stoichiometric, 2
ON/air-fuel ratio around peak
3-4 ON/10 kPa
Compression ratio
5 ON/compression ratio
Coolant temperature
1 ON/10 K
2
00-
LU
15
WOT
I
4RL Ref1) [66.J
0-
LU
c-_
Ref (2) (497]
[43]
4
5
Z
-A12A [309] [SWEPT VOLUME)
A1
[246]
c
1
05
6
8
10
12
1
14
16
18
20
COMPRESSION RATIO
Figure 1.13 - Brake efficiency improvement with compression ratio for engines of
several swept volumes [171.
40
1.6 OBJECTIVES
The objective of this work is to contribute to the knowledge required to optimize the
configuration of an SI engine to properly balance tradeoff between mid-load efficiency
and low speed maximum torque. The main factor affecting this tradeoff has traditionally
been compression ratio. More recently, intake pressure boosting has also become a
significant factor, with reduced compression ratios being necessary for satisfactory
boosted engine performance at the knock limit. The development by MIT of an on-board
fuel reformer that has been shown to lower the octane number requirement of the primary
fuel [14] has introduced an additional factor. Five main tasks were carried out:
1.
Investigate the efficiency effects of increasing the compression ratio at several
load and air-fuel ratio conditions.
2. Investigate the relationship between spark retard and torque reduction.
3. Investigate the trends in knock limited spark timing with changes in compression
ratio, intake boost, reformate fraction, air-fuel ratio, and fuel type.
4. Refine a previously developed knock modeling methodology and investigate its
accuracy and sensitivity to initial conditions.
5. Apply experimental data to the optimization of engine performance and
efficiency.
41
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42
CHAPTER 2. EXPERIMENTAL METHOD
To explore the effects of compression ratio, operating conditions, and fuel reformate on
engine efficiency and knock limited performance, a series of low to medium load
efficiency experiments and a series of high load knock experiments were performed. The
set of experiments was repeated for each of three compression ratios. The experimental
methods and test matrices were designed to define as many trends as possible without a
prohibitively large number of test runs.
2.1 ENGINE SETUP
The test cell contains a single cylinder research engine, a dynamometer, fuel and air
supply systems, and related support equipment.
2.1.1 Engine and Dynamometer Specifications
The test engine is a Ricardo MK III single cylinder research engine retrofitted with a
Volvo B5254 pent-roof, 4-valve, central spark plug cylinder head. A tumble/swirl
inducing plate between the inlet manifold and the head is used at low and medium loads
to increase charge turbulence. The engine specifications are shown in Table 2.1.
Table 2.1 - Test Engine Specifications
Bore (mm)
Stroke (mm)
Displacement Volume (cm 2 )
Connecting Rod Length (mm)
Piston 1 Clearance Vol. (cm 2)/Comp. Ratio
Piston 2 Clearance Vol. (cm 2)/Comp. Ratio
83
90
487
158
55 / 9.8:1
46 / 11.6:1
Piston 3 Clearance Vol. (cm 2)/Comp. Ratio
39 /13.4:1
Valve
Ia Timing
Tii
IVO 00 BTC, IVC 600 ABC
EVO 480 BBC, EVC 120 ATC
Three custom-made forged pistons were used for the experiments, one for each
compression ratio. The baseline piston is similar to the stock Volvo piston, with the same
crown shape and first land height. The two high rc pistons have partially raised piston
crown "pop-ups" to decrease the combustion chamber clearance volume. Figure 2.1
43
shows a photograph of the three pistons and Appendix A shows the dimensions of the
high r. piston crowns. Figure 2.2 is an illustration of cross sections of the combustion
chamber at the top center position for the 13.4:1 piston.
Figure 2.1 - Base, medium, and high compression ratio pistons (left to right)
Figure 2.2 - Illustration of the 13.4:1 piston and cross sections of the combustion
chamber at top center piston position
An EATON 6000 series 50HP electric dynamometer is connected to the engine by a drive
shaft. This dynamometer is able to motor the engine or absorb power from it and
automatically adjusts its torque output to maintain a constant speed at a manual set-point.
44
2.1.2 Air and Fuel Supply Systems
The engine intake system, pictured in Figure 2.3, can be switched between two
configurations - ambient or boosted. In the ambient configuration the air is drawn
through a filter from the test cell. In the boosted configuration air is supplied by a
compressor in a separate test cell, then is filtered and regulated to the desired engine inlet
pressure. In both cases the air then goes through a laminar flow element and a pulsationdamping tank before being throttled into the engine intake.
regulator to
amiet
ompresso
a
filter
0-brggvalve
4bar
gage
laminar flow
Figure 2.3 -Intake air system
The engine liquid fuel system, outlined in Figure 2.4, can be configured to draw out of
either of two fuel tanks, and can be purged with nitrogen gas. During normal operation
the pump draws the fuel out of the main fuel tank. From the pump, the fuel flows
through the filter to the engine. At the engine, a regulator keeps the pressure in the fuel
line at a constant differential (approximately 3 bar) from the average intake manifold
pressure. Excess fuel that is not consumed by the injector is returned to the fuel tank by
the regulator. For high accuracy fuel flow measurements the fuel supply and return can
be switched to the second fuel tank, which is equipped with a balance. The nitrogen
purge system allows the fuel system to be cleared of fuel, except for the small volumes of
the accumulator, the regulator, and the injector supply line.
Figure 2.5 shows the gaseous fuel system used to supply simulated plasmatron gas to the
engine. A regulator is used to control the pressure upstream of the critical flow orifice.
45
o
leslo
t
p
regulator to
1bar gage
valve3
to fuel injector
slop
tank 1I
N2
selector
fuel
tank1I
from intake manifold
fle
hose
pump
------------valvel
valve5
differential
backpressure
regulator (~3bar)
valve2
fuel
valve4
accumulator
tank 2
balance
............
return
Figure 2.4 - Liquid fuel system
adjustable
regulator
solenoid
valve
critical flow
orifice
flame
arrestor
to intake manifold
H2,
co,
N2
Figure 2.5 - Gaseous fuel system (simulated plasmatron gas supply)
2.2 ENGINE CONTROL AND MEASUREMENTS
A diagram of transducers and gauges is shown in Figure 2.6. Descriptions are listed in
Table 2.2. The rest of this section describes the control and measurement system in
further detail.
2.2.1 Engine Control Unit
A MoTeC M4 engine controller is used to control the injector and ignition system. The
injection timing is set to 3850 BTC and the dwell is set between 4 ms and 8 ms,
46
depending on the compression ratio of the engine (high dwell for high re). The injector
pulse width and spark timing can be adjusted while the engine is running.
selector
fuel
tank 1
hose
...........................
funel
tank 2
TT2 PT1
throttle P
damping
from air supply
laminar flow
element
inj.
from gaseous fuel supply ..............
P2critical flow
orifice
P
P
P
dynamometerEN
heated oil supply
heatedlcooled coolant supply
exhaust to trench
Figure 2.6 - Engine control and measurement diagram
2.2.2 Fuel Flow Measurement
There are two methods of liquid fuel flow measurement. For fast measurements with
limited accuracy (+/- 2%) the fuel injector pulse width is used to determine the amount of
fuel injected for each cycle. The fuel pressure regulator keeps a constant differential
between the average intake manifold pressure and the fuel injector supply pressure. Thus
the flow through the injector orifice is constant when the injector is open. Using an
experimental calibration, the mass of fuel injected can be calculated from the injector
pulse width. The accuracy of this method is limited because of fuel temperature-density
effects, and because high-speed pressure fluctuations in the intake manifold vary the
pressure across the injector at the time of injection as operating conditions vary. For
more accurate measurements (+/- 0.5%) a "pail and scale" method is used. For this
47
method the fuel supply and return are switched to a container on a balance. The mass of
fuel consumed by the engine over a set time period is calculated by subtracting the final
mass of the container from the initial mass. The fuel mass flow rate is the mass of fuel
consumed divided by the time period.
Table 2.2 - Summary of experimental engine transducers and gauges
Name
CP1
CP2
CP3
CP4
DPT
ENC
Measurement
crankshaft position
crankshaft position
camshaft position
camshaft position
differential pressure transducer
crankshaft position encoder
02T
P1
P2
PTl
PT2
PT3
oxygen sensor
oil pressure gauge
gaseous fuel pressure gauge
air supply absolute pressure transducer
manifold absolute pressure transducer
cylinder pressure transducer
TT1
TT2
TT3
TT4
TT5
WB
engine coolant inlet thermocouple
air inlet thermocouple
exhaust thermocouple
engine coolant inlet thermocouple
engine oil inlet thermocouple
balance
Details
for engine speed display
for engine control unit
for pressure acquisition system
for engine control unit
for air volume flow rate display
360 pulses per revolution, for pressure
acquisition system
for lambda meter/display
for air supply pressure display
for MAP display
Kistler pressure transducer, converted to
voltage signal by charge amplifier,
measured by computer DAQ card
for coolant temperature display
for inlet air temperature display
for exhaust temperature display
for engine coolant heater thermostat
for engine oil heater thermostat/display
displays mass of fuel tank 2
A critical flow orifice is used to control the gaseous fuel flow rate. So long as the
absolute gas pressure upstream of the orifice is kept higher than double the manifold
pressure, flow through the orifice will be choked. For choked flow, the mass flow rate
depends only upon the orifice dimensions, the upstream pressure and temperature, and
the specific heat and molecular weight of the gas. The orifice calibration (performed
with air) is modified to take into account the reformate gas properties. The temperature
upstream of the orifice is relatively constant at room temperature due to the long flowpath through a copper tube. Thus, adjusting the upstream gas pressure with a regulator
48
sets the orifice flow rate to a known value. Since the range of each orifice is limited,
several orifices are required to achieve the required range.
2.2.3 Intake Pressure Measurement and Control
Absolute pressure transducers measure the intake pressure in two locations. One is after
the air flow meter but before the throttle plate (air tank pressure), and the second is
between the throttle and the engine (Manifold Absolute Pressure - MAP). Digital
readouts display the results. The manifold pressure is controlled by manually adjusting
the angle of a throttle valve that is driven by a stepper motor.
2.2.4 Temperature Measurement and Control
There are four temperature measurement points on the engine:
" Engine coolant inlet
" Intake air, in the tank between the flow meter and the throttle
" Exhaust, about 2 cm from the exhaust port outlet
" Engine oil inlet
The engine coolant temperature is controlled to 900 C +/- 20 C by an electric heater
connected to an electronic thermostat, and a cold-water heat exchanger connected to a
mechanical thermostat. An electric heater connected to an electronic thermostat controls
the oil temperature to approximately 700 C.
2.2.5 Air Flow Measurement
Air volume flow rate is obtained by measuring the differential pressure across a laminar
flow element. The volume flow rate is converted to a mass flow rate by using the ideal
gas law with the air tank pressure and temperature measurements. The air mass flow rate
is corrected for water content using a humidity measurement. The accuracy of the air
mass flow measurement is approximately +/- 2%.
2.2.6 Air-Fuel Ratio Measurement
A Universal Exhaust Gas Oxygen (UEGO) sensor measures the oxygen content of the
exhaust gas. A Horiba Mexa-11 O analyzer interprets the signal and displays the air-fuel
equivalence ratio (2).
49
2.2.7 Cylinder Pressure Measurement
A Kistler 6125A piezoelectric pressure transducer equipped with a flame arrestor
measures the cylinder pressure. A charge amplifier converts the current signal from the
transducer to a voltage signal. The voltage signal is measured once every crank degree
by a National Instruments 6023E data acquisition card triggered by a BEI crankshaft
encoder. A program written in National Instruments LabVIEW records 300 cycles of
cylinder pressure data when triggered by the user.
Burn rate analysis software is used to calculate several quantities from the acquired
cylinder pressure data. They include:
0
NIMEP (Net Indicated Mean Effective Pressure)
0
COV (Coefficient of Variation) of NIMEP
0
GIMEP (Gross Indicated Mean Effective Pressure)
0
COV of GIMEP
0
Peak pressure
0
Crank angle of peak pressure
*
0-10% mass fraction burned time
*
0-50% mass fraction burned time
0
10-90% mass fraction burned time
A Visual Basic macro automatically runs the burn rate analysis software with inputs
exported from the data collection spreadsheet and writes the results to the spreadsheet.
2.2.8 Knock Detection
An audible knock method was used to detect knock for this work. An equalizer set to
remove all frequencies except those near 6 kHz and 12 kHz filtered the signal from a
microphone placed approximately 1 cm above the valve cover. Headphones were used to
listen to the resulting audio signal and identify knock. This method was selected because
it was shown in previous work to be consistent [8], because it closely matches the method
used during real engine calibration, and because it avoided the complexity of a sensorbased knock detection system.
50
2.3 CHANGING COMPRESSION RATIO
Three pistons were used during the experiments. When the first piston was installed in
the engine, a new cylinder liner, new rings, and new connecting rod bearings were also
installed. The engine was then broken in for 50 hours at varying speeds and loads.
Motoring and firing cylinder pressure data was recorded and analyzed to ensure proper
sealing and proper operation of the data acquisition system. Efficiency and emissions
data was recorded and compared to the stock Volvo piston to ensure consistency.
To minimize variation in ring pack performance and eliminate the need for cylinder liner
honing, the first and second compression rings were re-used for the second and third
piston. After each piston was installed, the engine was broken in for 20 hours and
cylinder pressure data was recorded under motoring and firing conditions to ensure
proper sealing and proper operation of the data acquisition system.
2.4 EFFICIENCY EXPERIMENTS
A series of tests were performed to determine trends in indicated efficiency and other
engine parameters with changes in operating conditions and compression ratio.
2.4.1 Experimental Procedure
Before each set of data is taken, the engine is fully warmed up. The procedure for each
data point is as follows:
1.
Set engine speed in the dynamometer controller.
2. Make an initial estimate of MBT (Maximum Brake Torque) timing from previous
trends.
3. Adjust fueling and air flow to achieve the desired air-fuel ratio and NIMEP (from
a real-time readout).
4. Perform a timing sweep while recording and processing cylinder pressure data to
identify MBT timing to within 10 CA. Re-check air-fuel ratio and NIMEP.
5. At MBT timing, record operating conditions and cylinder pressure data.
6. Keeping control parameters constant, record the change in fuel tank mass (from
the balance) over a 600 sec (1500 rpm) or 420 sec (2500 rpm) time period.
51
The experimental results are recorded in a spreadsheet and used to calculate indicated
efficiencies, mass fraction burned angles, and other useful parameters. The k
measurement from the UEGO sensor and the ), calculated from the measured air and fuel
flow rates are compared to ensure consistency.
2.4.2 Fuel
The fuel used for the efficiency experiments, termed "toliso", is a mixture of 70% isooctane with 6.0 mL of TEL per gallon (PRF RON of 120) and 30% toluene. A high-ON
fuel was required to avoid knock at high loads and high compression ratios. This mixture
was selected because it has a similar alkane/aromatic ratio, H/C ratio, energy content, and
specific gravity to gasoline. The high vaporization temperature was not considered to be
important since all experiments were performed under fully warmed conditions.
2.4.3 Operating Conditions and Compression Ratio
The test matrix for the efficiency experiments is shown in Figure 2.7. At each rc and each
engine speed an air-fuel ratio sweep was performed and load sweeps at two air-fuel ratios
were performed for a total of 42 data points.
Rc
9.8:1
11.6:1
1=1.3
X=1.0
2bar
NIMEP
2bar
NIMEP
2500rpm
X=1.0
X=1.3
X=1.6
1500rpm
4bar
NIMEP
4bar
NIMEP
4bar
NIMEP
X=1.0
8bar
NIMEP
X=1.3
Speed
1500rpm
2500rpm
13.4:1
13.4:1
1500rpm
160p
2500rpm
8bar
NIMEP
Figure 2.7 - Test matrix for efficiency experiments
2.5 KNOCK LIMITED MINIMUM SPARK RETARD EXPERIMENTS
A series of tests were performed to determine the relationship between spark timing and
engine output and to identify trends in knock limited spark timing with changes in
operating conditions and compression ratio.
52
2.5.1 Experimental Procedure
Before each set of data is taken, the engine is fully warmed up. Each data point consists
of a spark-timing sweep with a high-octane fuel - PRF 120 for a PRF data point and
toluene for a TRF data point (see Section 2.5.2 ) - and a Knock Limited Spark Advance
(KLSA) fuel octane sweep. The procedure for each data point is as follows:
Spark timing sweep
1.
Set the engine speed with the dynamometer controller.
2. Position the throttle plate in the wide-open position.
3. Make an initial estimate of MBT (Maximum Brake Torque) timing from previous
trends.
4. For naturally aspirated points, adjust the fuel injector pulse width to get a
stoichiometric air-fuel ratio. For boosted points, adjust the boost pressure
regulator and injector pulse width to achieve the desired NIMEP and air-fuel
ratio. For points with fuel reformate, the gaseous fuel flow rate must also be
adjusted to achieve the desired reformate fraction.
5. Set spark timing to approximately 40 CA earlier than estimated MBT timing. If
the spark timing cannot be advanced because the point is knock limited (high rc,
high boost points), set the spark timing as early as possible while avoiding the
onset of pressure oscillations. If the spark timing cannot be advanced because the
point is cylinder pressure limited (high re, high boost points), set the spark timing
as early as possible without exceeding the maximum cylinder pressure of 11 Obar.
6. Record operating conditions and cylinder pressure data.
7. Retard the spark timing in steps of 2' CA, repeating step 6 for every spark timing
until combustion becomes unstable or exhaust temperature approaches 7500 C. At
some timings spark discharge noise will interfere with the data acquisition system,
so step sizes must be adjusted by 0.5' CA. At retarded spark timings the throttle
must be closed slightly to maintain the correct exhaust air-fuel ratio.
Fuel octane sweep
1.
Keeping the engine motoring, flush the fuel system with nitrogen.
2. Load the fuel system with the highest ON fuel that will knock with spark timing
set to MBT.
53
3. Let engine stabilize at a spark timing that avoids knock for 2-3 minutes.
4. Keeping fuel flow and air-fuel ratio constant, advance spark timing by increments
of 10 CA until knocking cycles are clearly heard through the audio system.
5. Retard the spark timing by 1 CA and record operating conditions and cylinder
pressure data.
6. Keeping the engine motoring, flush the fuel system with nitrogen and load the
next lower ON fuel.
7. Repeat steps 3 - 6 until spark is retarded to the point that combustion is unstable,
exhaust temperature approaches 750 C, or runaway knock cannot be avoided.
For UTG fuels (see Section 2.5.2 ) the spark timing cannot be advanced to MBT during
the spark-timing sweep, so the procedure is modified as follows:
1.
Set the spark timing 4-6* CA later than the expected knock limit.
2. Since timing cannot be advanced to MBT due to knock, the required boost
pressure must be approximated by matching the NIMEP of the toluene sparktiming sweep under similar operating conditions. Adjust the boost pressure
regulator and injector pulse width to achieve the desired X and match the NIMEP
of the corresponding toluene data point. For points with reformate, the gaseous
fuel flow rate must also be adjusted to achieve the desired reformate fraction.
3. Retard spark timing until combustion becomes unstable or exhaust temperature
approaches 7500 C.
4. Keeping fuel flow and X constant, advance spark timing by increments of 1 or 20
CA until knocking cycles are clearly heard through the audio system. Record
operating conditions and cylinder pressure data at each spark timing.
The experimental results are recorded in a spreadsheet and used to calculate bum angles
and other relevant parameters. The X measurement from the UEGO sensor and the X
calculated from the measured air and fuel flow rates are compared to ensure consistency.
2.5.2 Fuels
Since gasoline is available only in a limited set of formulations, three types of fuels were
used in the experiments to explore the entire range of operating conditions and fuel
qualities of interest:
54
UTG
Unleaded test gasolines, commonly termed "indolene", are standardized fuels typically
used for emissions certification and laboratory work. For this work UTG91 and UTG96
were obtained from Chevron Phillips Chemical Company LLC to represent the
performance of gasoline. The relevant specifications of these fuels are in Table 2.3.
Table 2.3 - UTG fuel properties
Specification
RON
MON
Sensitivity
Alkanes,%vol
Aromatics,%vol
Olefins,%vol
H/C ratio
Typical
UTG91
90.8
83
7.8
70
24
6
1.89
Values
UTG96
96.1
87
9.1
67
28
5
1.86
PRF
Primary reference fuels are the reference for the ASTM octane rating methods. By
definition, the RON and MON of PRFs are equal. For ON of 0 to 100, PRFs are mixtures
of iso-octane and n-heptane, the octane numbers of which are the volume fraction of isooctane under standard conditions. For ON of 100 to 120, PRFs are mixtures of iso-octane
and TEL (tetraethyl lead), the proportions of which are specified in the ASTM 2699 and
ASTM 2700 standards by:
ON =100 +
28.28T
2
1+ 0.736T + 1+ 1.472T - 0.035216T
(2.1)
Where T is the amount of TEL in mL per gallon of iso-octane. For this work, PRFs were
mixed in increments of 5 ON. To avoid directly handling concentrated TEL, which is
toxic, fuels above PRF 100 were mixed by diluting PRF 120 with iso-octane to obtain the
required TEL concentration.
55
TRF
To better represent the sensitivity and aromatic content of gasolines, the knock
experiments were also performed with toluene reference fuels. TRFs are mixtures of
toluene and n-heptane. TRF RON and MON are plotted as a function of n-heptane
volume fraction in Figure 2.8. For this work TRFs were mixed in increments of 5 ON.
110 RON = -0.0097x 2 - 0.4419x + 111.29
+ RON
M MON
100 -
E 90 z
I.
0
0
80 70 60 MON
50
=
-O.0031x 2
-
0.7498x
+ 103.24
i
10
20
30
40
50
60
%v n- heptane (bulk toluene)
Figure 2.8 - RON and MON of toluene reference fuels as a function of volume
fraction n-heptane. Compiled by Shell Global Solutions [18] from ASTM vol 05.04,
Table 28 and internal data.
2.5.3 Operating Conditions and Compression Ratio
The test matrix for the knock experiments is shown in Figure 2.9. For each r" and each
fuel type an air-fuel ratio sweep was performed and boost sweeps at two air-fuel ratios
were performed. Reformate sweeps were performed from a boosted stoichiometric point
and a base lean point. Not all of the points were taken for the UTG fuels. In total there
are 83 data points.
56
Rc
Fuel Type
X=1.0*
WOT
PRF
9.8:19.1 - - - - - TRF
X=1.3*
I
PRF
11.6:1
TRF
UTG91*
UTG96***
X=1.0
200/a boost
(1.2X WOT
NIMEP)
X=1.3**
15% boost
(1.15X X=1.0
WOT NIMEP)
TRF
UTG96*
'match X= 1.0
WOT NIMEP
15% reformate
fraction
I__
PRF
13.4:1
=1.6*
bo osted to
mat ch X=1.0
WO T NIMEP
X=1.3*
boosted to
match X=1.0
OT NIMEP
400/ boost
(1.4X WOT
NIMEP)
)L=1.3**
300a boost
(1.3X X=1.0
WOT NIMEP)
X=1.3*
match 2=1.0
WOT NIMEP
3 00
A reformate
fraction
400h boost
15% reformate
fraction
k=1 .0
400a boost
3 00
/ reformate
fraction
Figure 2.9 - Test matrix for knock experiments
57
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58
CHAPTER 3. EXPERIMENTAL RESULTS
This chapter presents the data collected using the apparatus and procedures described in
Chapter 2. The experimental results are used to evaluate the effects of engine operating
parameters on efficiency and knock-limited performance. An extended data set is
available in the appendices.
3.1 EFFICIENCY RESULTS
The data presented in this section was collected as described in Section 2.4 . The effects
of air-fuel ratio and load on indicated efficiency are discussed first, followed by a
description of the efficiency effects of compression ratio.
3.1.1 Effects of Air-Fuel Ratio and Load
Figure 3.1 shows the experimental net indicated efficiency (rjin) for air-fuel equivalence
ratio (k) sweeps at the mid-load point of 4.0 bar NIMEP (Net Indicated Mean Effective
Pressure). Data is displayed for all three compression ratios at 1500 rpm and 2500 rpm.
The normalized improvement from the stoichiometric efficiency is shown in Figure 3.2.
The curves are drawn to illustrate the expected trends between data points, and match
previous work that found peak fin to be at k = 1.5 for a similar load, 1500 rpm, and 10:1
rc [6].
Figure 3.3 shows the effect of k on gross indicated efficiency (rtig) for the same data
points. The maximum increase in rl'ig is about two-thirds that of the maximum increase
in tifn, indicating that about two-thirds of the improvement from increasing air fuel ratio
is from increased y and decreased heat transfer. The remaining improvement is from
reduced pumping work.
At low k, the normalized increase of agin with k appears stronger for low rc than for high
rc. Figure 1.5 shows that the magnitude of the efficiency improvement from increasing y
(i.e. from increasing k) remains approximately constant as rc increases. Since efficiency
is increasing with compression ratio, at high rc the improvement from increasing y is
smaller as a fraction of the baseline efficiency.
At low rc, i1fig peaks and starts to decay at a lower air-fuel ratio than for high rc. The
extension of X for peak efficiency at high rc is likely due to reduced combustion durations
59
and cycle-to-cycle variability. The same trend is visible for rIjn, but decreasing pumping
work extends the peaks to higher air-fuel ratios. Figure 3.4 shows combustion event
timing at the 9.8:1 r, as a function of X. The crank angles of spark and 10%, 50%, and
90% mass fraction burned are shown. As the mixture becomes leaner the laminar flame
speed decreases, leading to longer combustion durations. Figure 3.5 shows combustion
event timing under lean conditions as a function of rc. As r, increases, the pressures and
temperatures in the combustion chamber increase, resulting in a higher laminar flame
speed, which is evident in the decreased 0-10% mass fraction burned times. The decrease
in 10-90% mass fraction burned time is probably due to a combination of the increased
laminar flame speed, and an increase in turbulence from the fluid being "squished"
towards the center of the chamber by the piston "pop-up".
As engine speed increases, there is less time for heat loss from the burned gases for each
cycle, so in all cases qfig at 2500 rpm is higher than at 1500 rpm. Net efficiency is also
higher at the higher speed, but by a lesser amount because the increased efficiency from
lower heat loss must be counteracted with increased throttling to maintain constant load,
resulting in increased pumping losses.
Figure 3.6 and Figure 3.7 show the increase in Tlin with increasing load for X of 1.0 and
1.3. The efficiency improvement is normalized around the mid-load 4 bar NIMEP point
in Figure 3.8 and Figure 3.9. At mid-load, net efficiency improves by about 7% per bar
NIMEP relative to the efficiency at 4 bar NIMEP. Although well aligned, the curves for
high r, are slightly steeper than for low rc. This is likely because the higher efficiency at
high r, reduces the intake manifold pressure necessary to maintain the same load. At the
reduced manifold pressures, the reduction in throttling losses from an incremental load
increase improves efficiency more than at higher manifold pressures. The improvement
in rjin with load for the stoichiometric air-fuel ratio is slightly better than for the lean airfuel ratio. Again, this is because manifold pressures are higher under lean conditions so
there is not as much benefit from reduced throttling.
The gross indicated efficiencies, shown in Figure 3.10 and Figure 3.11, increase with
load at approximately one quarter the rate of the net efficiencies. Since the difference
between rfin and t mig is due to pumping work, reduction in throttling is responsible for
about 75% of the improvement in agin as NIMEP increases. The rest of the improvement
is mostly due to heat loss becoming a smaller portion of the total charge energy.
60
3.1.2 Effects of Compression Ratio
Figure 3.12 shows the change of net indicated efficiency with compression ratio for a
range of X at mid-load. The efficiency increase is normalized in Figure 3.13. At the base
rc of 9.8:1, 11in improves by about 2.5% per unit rc. Efficiency appears to peak at a rc of
about 14:1 with an improvement of 6-7%. This agrees well with existing data [17]. The
improvement is better at 2500 rpm than at 1500 rpm, likely because the increased heat
transfer at high r0 does not have as much of an effect when the amount of time for heat
transfer is lower. There does not appear to be a clear trend with X.
Figure 3.14 and Figure 3.15 show the increase of Tgi. with increasing r, for a range of
loads at X of 1.0 and 1.3. The efficiency improvement from the base compression ratio is
normalized in Figure 3.16 and Figure 3.17. The improvement of jfin is better at higher
loads, where the increased heat transfer from increasing rc is a smaller fraction of the
overall charge energy. The experimental results also show that load has a greater effect
on the relationship between re and ifi under lean conditions than it does at , = 1.0.
0.36
-+- Rc = 9.8:1, 1500 rpm
-- Rc = 9.8:1, 2500 rpm
-- Rc = 11.6:1, 1500 rpm
0.35 '
-z-Rc = 11.6:1, 2500 rpm
-- Rc = 13.4:1, 1500 rpm
-e-Rc = 13.4:1, 2500 rpm
0.34-
4 0.33 0.32-
z 0.31 0.30.290.9
1
1.1
1.2
1.3
1.4
1.5
1.6
1.7
Lambda
Figure 3.1 - Change of net indicated efficiency with lambda; 4.Obar NIMEP.
61
12
-+- Rc =
Rc =
-Rc =
-A- Rc =
-4-Rc =
S10-
9.8:1, 1500 rpm
9.8:1, 2500 rpm
11.6:1, 1500 rpm
11.6:1, 2500 rpm
13.4:1, 1500 rpm
-e-- Rc = 13.4:1, 2500 rpm
C
S-
2-
Z
40)-
0.9
1.1
1
1.3
1.2
1.4
1.5
1.6
1.7
Lambda
Figure 3.2 - Normalized change of net indicated efficiency with lambda; 4.Obar
NIMEP. 4)
0.4--+
Rc = 9.8:1, 1500 rpm
-- Rc = 9.8:1, 2500 rpm
0.39 --
Rc =
Rc =
-- Rc =
-e-Rc =
>-%
S0.38 -
.9
11.8:1,
11.6:1,
13.4:1,
13.4:1,
u'0.37 0.36-
0.35 0.34 0.330.9
1
1.1
1.2
1.3 1.4
Lambda
1.5
1.6
1.7
Figure 3.3 - Change of gross efficiency with lambda; 4.bbar NIMEP.
62
1500
2500
1500
2500
rpm
rpm
rpm
rpm
30
-+-
..........
Spark, 1500 rpm
-m-10% mfb, 1500 rpm
20-
mfb, 1500 rpm
-0-90% mfb, 1500 rpm
-&-50%
10-
-*- Spark, 2500 rpm
-e- 10% mfb, 2500 rpm
-A- 50% mfb, 2500 rpm
0-10-
-e- 90% mfb, 2500 rpm
(-20-30-40-50 -
-60
1.1
0.9
1.7
1.5
1.3
Lambda
Figure 3.4 - Combustion event timing; 4.Obar NIMEP, r, = 9.8:1.
30
20
-
10
-
-4-$Spark, 1500rpm
-U-10% mfb, 1500 rpm
-+-50% mfb, 1500 rpm
+90% mfb, 1500 rpm
-+- Spark, 2500 rpm
-E- 10% mfb, 2500 rpm
10
0 0-
-A- 50% mfb, 2500 rpm
-
-e- 90% mfb, 2500 rpm
(-20-30-40-50-60
I
9
10
I
11
I
I
12
13
14
RF
Figure 3.5 - Combustion event timing; 4.Obar NIMEP, X = 1.6.
63
0.38
Rc = 98:1 1500 rpm
-e- Rc = 9.8:1, 2500 rpm
--
0.36 -
-- Rc =
-r-Rc =
-- Rc =
-e- Rc =
0.34-
11.6:1,
11.6:1,
13.4:1,
13.4:1,
1500
2500
1500
2500
rpm
rpm
rpm
rpm
lw 0.32 o
0.3 -
z 0.280.26 -
0.24 -
i
i
i
i
1
2
3
4
5
6
7
8
9
NIMEP (bar)
Figure 3.6 - Change of net indicated efficiency with NIMEP; X = 1.0.
0.4
-+-Rc = 9.8:1, 1500 rpm
0.38 -
-+-Rc = 11.6:1, 1500 rpm
--A- Rc = 11.6:1, 2500 rpm
0.36 -
-4- Rc = 13.4:1, 1500 rpm
-e-Rc = 13.4:1, 2500 rpm
---
Rc = 9.8:1, 2500 rpm
w 0.34V
.- *
U
0.32 -
0.30.28-
0.26
1
2
3
4
5
6
7
8
9
NIMEP (bar)
Figure 3.7 - Change of net indicated efficiency with NIMEP; X
64
=
1.3.
20
8
-
-*- Rc = 9.8:1, 1500 rpm
?15 -
~+-Rc = 9.8:1, 2500 rpm
-Rc = 11.6:1, 1500 rpm
10-
-- Rc = 11.6:1, 2500 rpm
-4--Rc = 13.4:1, 1500 rpm
-e-Rc = 13.4:1, 2500 rpm
-U
Z -10-
& -15
0 -20-
-25
1
2
I
3
I
5
I
4
I
6
7
I
8
9
NIMEP (bar)
Figure 3.8 - Normalized change of net indicated efficiency with NIMEP; X
20 -- --
=
1.0.
-+--Rc = 9.8:1, 1500 rpm
615
+ Rc = 9.8:1, 2500 rpm
-+- Rc = 11.6:1, 1500 rpm
S10Z
--A- Rc = 11.6:1, 2500 rpm
-0-Rc = 13.4:1, 1500 rpm
-e-Rc = 13.4:1, 2500 rpm
-
z -10C
CM -15S-20-25-
I
1
2
3
4
5
6
7
8
9
NIMEP (bar)
Figure 3.9 - Normalized change of net indicated efficiency with NIMEP; X
=
1.3.
65
0.4
-+-Rc =9.8:1, 1500 rpm
+R %= .a8 : , 2500 rpml
-
0.38 -
Rc = 11.6:1, 1500 rpm
,A- Rc = 11.6:1, 2500 rpm
-4-Rc = 13.4:1, 1500 rpm
-e-Rc = 13.4:1, 2500 rpm
0.36-
0.34-
0.32-
0.3 1
2
3
I
4
I
5
I
6
I
7
8
9
NIMEP (bar)
Figure 3.10 - Change of gross indicated efficiency with NIMEP; X = 1.0.
0.42
-+- Rc = 9.8:1, 1500 rpm
-e- Rc = 9.8:1, 2500 rpm
-0.- Rc = 11.6:1, 1500 rpm
--A- Rc = 11.6:1, 2500 rpm
-4-Rc = 13.4:1, 1500 rpm
-e--Rc = 13.4:1, 2500 rpm
W 0.38 -)
0.36 0
0.34-
0.32 1
2
3
4
5
I
6
I
7
I
8
9
NIMEP (bar)
Figure 3.11 - Change of gross indicated efficiency with NIMEP; X
66
=
1.3.
n0
-+-Lambda
-s-Lambda
Lambda
-- Lambda
0.35 > 0.34
-
= 1.0, 1500 rpm
= 1.0, 2500 rpm
= 1.3, 1500 rpm
= 1.3, 2500 rpm
-*--Lambda = 1.6,1500 rpm
-e- Lambda = 1.6, 2500 rpm
0.320.31 Z
0.3 -
0.29
0.28
-
'
i
9
10
11
12
13
14
Figure 3.12 - Change of net indicated efficiency with compression ratio for a range
of X; 4.Obar NIMEP.
8
--
7
--4-e-
Lambda = 1.0, 1500 rpm
-+- Lambda = 1.0, 2500 rpm
6
W. 5
Z
Lambda =
Lambda =
Lambda =
Lambda =
1.3,
1.3,
1.6,
1.6,
1500
2500
1500
2500
rpm
rpm
rpm
rpm
3
2
0
T
9
-
10
11
12
13
14
Figure 3.13 - Normalized change of net indicated efficiency with compression ratio
for a range of X; 4.Obar NIMEP.
67
0.38
-
-+-NIMEP = 2.0 bar, 1500 rpm
0.36 -
-+-
NIMEP = 2.0 bar, 2500 rpm
- NIMEP = 4.0 bar, 1500 rpm
0.34
-
--
-*-NIMEP = 8.0 bar, 1500 rpm
e- NIMEP = 8.0 bar, 2500 rpm
0.32-
.
0.3-
W
'
z
NIMEP = 4.0 bar, 2500 rpm
0.28
-
0.26
-
0.24
-
0.22 -
:
0.2
9
10
11
12
13
14
Figure 3.14 - Change in net indicated efficiency with compression ratio for a range
of loads; X=1.0.
0.4-
--
0.38
-
0.36
-
-,- NIMEP = 4.0 bar, 1500 rpm
-A-
0.34-
.
W
0.32
NIMEP = 2.0 bar, 1500 rpm
-+- NIMEP = 2.0 bar, 2500 rpm
NIMEP = 4.0 bar, 2500 rpm
-0- NIMEP = 8.0 bar, 1500 rpm
-e- NIMEP = 8.0 bar, 2500 rpm
-
0.3
0.28-
0.26
-
0.24
-
0.22
i
9
10
11
12
13
14
Figure 3.15 - Change in net indicated efficiency with compression ratio for a range
of loads; X =1.3.
68
8
NIMEP = 2.0 bar, 1500
-+- NIMEP = 2.0 bar, 2500
-A-NIMEP = 4.0 bar, 1500
,h- NIMEP = 4.0 bar, 2500
--
S6
0-l
0 7
-0--NIMEP = 8.0 bar, 1500 rpm
-e- NIMEP = 8.0 bar, 2500 rpm
W. 5
cc
z
rpm
rpm
rpm
rpm
(4
S3
4)2
I1
a)
cc
0 9
- 1
9
10
11
12
13
14
Figure 3.16 - Normalized change in net indicated efficiency with compression ratio
for a range of loads; X=1.0.
8
-+-NIMEP = 2.0 bar, 1500 rpm
--
NIMEP = 2.0 bar, 2500 rpm
-h-NIMEP = 4.0 bar, 1500 rpm
-A- NIMEP = 4.0 bar, 2500 rpm
06
-*- NIMEP = 8.0 bar, 1500 rpm
-e- NIMEP = 8.0 bar, 2500 rpm
S
W5
3
z
C
S2
C
0 I
9
10
1
12
11
12
13
14
Figure 3.17 - Normalized change in net indicated efficiency with compression ratio
for a range of loads; X=1.3.
69
3.2 KNOCK LIMITED PERFORMANCE RESULTS
This section begins with a description of the relationship between spark timing and
engine torque output. The effects of varying air-fuel ratio, boost pressure, fuel reformate
fraction, and compression ratio on knock limited engine performance are subsequently
presented. Data was collected according to the procedure described in Section 2.5
3.2.1 Ignition Timing
The change of NIMEP with spark timing for a range of X with PRF 120 and toluene fuel
is shown in Figure 3.18. Spark timing for maximum NIMEP (MBT timing) advances
with increased X due to a slower flame speed. MBT timing is also more advanced for
PRF 120 than for toluene. These results agree with the findings of [19] which show that
TRF has similar combustion duration to gasoline while PRF is slightly slower. NIMEP
drops off as spark timing is retarded from the optimum because later combustion
decreases the volume ratio, and consequently temperature ratio, through which the
burned gases are expanded, resulting in less work extraction.
Figure 3.19 shows the same data as Figure 3.18, except with NIMEP normalized to
maximum NIMEP, and with spark timing normalized to MBT timing. The normalized
spark timing, termed "spark retard", is the spark timing, in 'ATC, minus the MBT spark
timing. It represents the number of crank degrees that the spark timing has been shifted
from that for maximum torque. It is evident from this chart that NIMEP of the test points
with slower combustion (i.e. with more advanced MBT timing) does not decrease as
quickly with spark retard as it does for test points with faster combustion. This trend can
also be seen in Figure 3.20, where the NIMEP of the low r, test points, which are slower
burning, drops more slowly with spark retard than for the high r, test points.
Another timing parameter, termed "combustion retard", has been developed to better
describe the change in combustion phasing from the optimal. Combustion retard is the
location of 50% mass fraction burned, in 'ATC, minus the location of 50% mass fraction
burned for MBT spark timing. It represents the number of crank degrees that the center
of the combustion event has been shifted from the timing for maximum torque. Figure
3.21 shows the relationship between spark retard and combustion retard for a range of X
with PRF120 and toluene fuel. It can be seen that the timing of 50% mass fraction
burned reacts differently to spark timing as operating conditions change. Figure 3.22
70
shows the change of normalized NIMEP with spark retard for a range of compression
ratios, fuels, air-fuel ratios, and intake pressures. There is significant spread, but when
combustion retard is substituted for spark retard, as shown in Figure 3.23, all of the points
fit well to a single diagonally asymptotic curve. The equation of the curve fit is:
NIMEP
NIMEP
=I - 0.168
-
+ 4.443
\ 10-3(50%mjb -
2\0.s431~~
050%mfb,MBT
1]
(3.1)
NIMEPMBT
where
0
50%m.t
is the crank angle of 50% mass fraction burned in "ATC and 0
50%mfb,MBT
the crank angle of 50% mass fraction burned at MBT spark timing. The quantity
0
50%mfb,MBT
0
is
50%mf-
is the combustion retard.
Under near-knocking conditions, the relationship between spark timing and location of
50% mass fraction burned is affected by early chemical heat release. Figure 3.24 shows
the heat release profiles averaged over 300 cycles for several spark timings. The solid
lines are from data taken using toluene as a fuel. The dashed lines are from data taken
using TRFs with octane numbers that yield close-to-knocking conditions (i.e. advancing
the spark timing by one degree would produce audible knock). As the spark timing
becomes more retarded and fuel octane number is decreased to maintain near-knocking
conditions, early (non-flame) heat release becomes more significant and flame speed
appears to increase. Under severely retarded conditions there are two distinct periods of
heat release - one from spontaneous reactions in the unburned gas, which peaks near top
center crank position, and a later one from the propagation of the turbulent flame. These
reactions, although they do not proceed to full autoignition and knock does not occur,
result in a shortening of the burn durations. Figure 3.25 shows that the reduced burn
duration causes 50% mass fraction burned to occur up to 4 degrees earlier for the same
spark timing.
Since the time of 50% mass fraction burned gives a better indication of when combustion
is occurring, and since it correlates better to torque loss, combustion retard is used in the
knock experiments as the indicator of combustion phasing.
3.2.2 Effects of Air-Fuel Ratio
Figure 3.26 and Figure 3.27 display the combustion retard required to just avoid knock as
a function of fuel RON and air-fuel ratio for PRFs and TRFs respectively. The
stoichiometric data points were taken at WOT (Wide-Open Throttle) with no boost. Air71
fuel ratio was increased from stoichiometric by boosting inlet pressure to match NIMEP
at MBT timing to the stoichiometric WOT case. High-octane fuel (PRF 120 for PRF
experiments, toluene for TRF experiments) was used when adjusting air-fuel ratio to
ensure that timing was not knock-limited. The data is also available as 2D charts for each
compression ratio and fuel type in Appendix B. Figure 3.28 shows the change in
combustion retard for UTG91 and UTG96 with k at a compression ratio of 11.6:1, with
reference fuel data plotted for comparison. There are distinct differences between the
results for PRFs, for which combustion retard increases with k, and the results for TRFs,
for which combustion retard decreases with k. The trends for the UTG fuels are roughly
half way between the PRFs and the TRFs, which results in little net change with k.
Figure 3.29, Figure 3.30, and Figure 3.31 show the normalized change in the combustion
retard required to just avoid knock with changing k for the three compression ratios and a
range of fuels. These charts show more clearly that PRFs require increasing combustion
retard with increasing k. The increase is mild at the 9.8:1 re, which agrees with previous
findings [8], and becomes more severe at high rc. The increase is also more severe for the
change from k of 1.3 to 1.6 than from 1.0 to 1.3. At all compression ratios the trend for
TRFs is a mild decrease in the amount of combustion retard required with increasing k.
Again, it is apparent that the trend for UTG fuels is roughly in between the two types of
reference fuels.
To increase X in these experiments the inlet pressure was boosted, decreasing the ratio
between the inlet pressure and the peak pressure. A possible reason why PRFs require
more spark retard than TRFs is that the lower specific heat of TRFs causes peak end-gas
temperatures to decrease more with the decreased pressure ratio. This effect is discussed
in more detail in Section 4.4.2 .
3.2.3 Effects of Boost
The boost sweeps were performed by setting MAP for the required amount of boost, then
successively lowering the octane number of the fuel and measuring the amount of
combustion retard required to avoid knock. Figure 3.32 and Figure 3.33 show a sample
of the results for PRFs and TRFs respectively at k = 1.0 and r, = 11.6:1. Lines are drawn
through points with the same fuel mixture. Figure 3.34 shows the same TRF data, except
with the axes swapped and with lines of constant MAP drawn using the combustion
retard-NIMEP relationship from Eq. ( 3.1). The vertical distance between the point on
72
the curve and the horizontal line at NIMEPMBT represents the torque lost from retarding
spark to avoid knock. As MAP is increased, NIMEPMBT increases, but torque loss also
increases.
The complete sets of PRF and TRF combustion retard data for stoichiometric boosting
are plotted in Figure 3.35 and Figure 3.36 respectively. The remaining 2D plots of
stoichiometric boosting data and results for boosting at k of 1.3 are available in Appendix
C. For the range of boost levels and fuel ON investigated, NIMEP always increases with
MAP, indicating that the torque gained from increasing the manifold pressure is greater
than the torque lost from retarding combustion.
PRFs consistently require more combustion retard than TRFs as NIMEP is increased.
One explanation is that for these boosting experiments, the peak pressure rises
approximately proportionally to MAP. Since the pressure ratio through which the endgas is compressed does not change significantly, there is no significant change in
compressed end-gas temperature. Simulated end-gas temperature profiles for boosted
and non-boosted cases are discussed in Section 4.4.2 . Autoignition reaction rates for
PRFs are more sensitive to pressure than for TRFs [20], so PRFs require more
combustion retard to avoid knock.
Figure 3.37 shows a comparison of UTG96 to PRF95 and TRF95 for boosting at a k =
1.3 and rc = 11.6:1. As was seen with the reaction of UTG fuels to increases in air-fuel
ratio, UTG96 behaves approximately half way between PRF and TRF. This trend is also
visible in Figure 3.38 and Figure 3.39, which show the increase of combustion retard
with boosted airflow and boosted NIMEP respectively for almost all of the test points
considered. (Points with very high combustion retard - greater than 300 - or with spark
timing earlier than MBT were omitted.) Although there is some spread to the data, it is
clear that PRFs, which require about 5' CA of combustion retard per bar NIMEP, need
about three times as much combustion retard as TRFs when boosted to achieve the same
NIMEP.
3.2.4 Effects of Plasmatron Reformate Addition
Fuel reformate was added under two different operating conditions for each compression
ratio. The first condition was stoichiometric with 40% boost, for which inlet pressure
was boosted so that NIMEP at MBT spark timing was 1.4 times NIMEPMBT at unboosted
WOT. The second condition was at
=1 .3 with MAP boosted to match NIMEPMBT at
73
unboosted stoichiometric WOT. In all cases the boost pressure was set while operating
with high ON fuel (PRF 120 for PRF experiments, toluene for TRF experiments) to
ensure that timing was not knock-limited. The reformed fraction is defined as the mass
of the fuel in the gaseous reformate, which is the mass of the carbon and hydrogen
constituents, divided by the total mass of the liquid and reformed fuel.
Figure 3.40 and Figure 3.41 show knock limited combustion retard and NIMEP for
reformate addition to PRFs and TRFs respectively. Data for boosted stoichiometric
operation at rc of 11.6:1 are shown. Data for other operating conditions and compression
ratios can be found in Appendix D. As reformate is added, autoignition reactions are
slowed and combustion retard can be decreased, resulting in increased NIMEP. Figure
3.42 shows how reformate addition can be used to recover torque lost from retarding
timing to avoid knock with increased boost. As boost is applied NIMEP increases, but
this increase is moderated because spark timing must be moved away from MBT timing
to avoid knock. Then, as reformate is added, spark timing can be advanced and NIMEP
increases further according to the relationship from Eq. ( 3.1). Reformate can also be
used to recover torque lost from spark retard as compression ratio is increased, as shown
in Figure 3.43.
Figure 3.44 shows a comparison of reformate addition to UTG fuels, which represent
gasoline, as compared to PRFs and TRFs at X = 1.3 and r, = 11.6:1. The results indicate
that the decrease of combustion retard for gasolines with increased reformed fraction is
roughly halfway between that of TRFs and PRFs with similar ON.
The complete set of data for adding reformate at the stoichiometric boosted condition to
PRFs and TRFs are shown as 3D surfaces in Figure 3.45 and Figure 3.46 respectively.
The decrease of combustion retard with reformate addition for each fuel is normalized in
Figure 3.47. Reformate addition is most effective when applied to PRFs, which are
composed entirely of alkane hydrocarbons. The combustion retard required to just avoid
knock decreases by about 2' CA per 3% reformed fraction. When applied to TRF fuels
with octane numbers of 95 or lower, which are composed of more than 20% n-heptane,
reformate addition is only slightly less effective than it is for PRF fuels. As the alkane
content decreases and aromatic content increases, reformate addition appears to become
less effective for TRFs to the point that there is no benefit for reformate addition to pure
toluene.
74
Autoignition chemistry modeling, described in Section 4.5 , indicates that the hydrogen
in the reformate slows down autoignition by converting hydroxy radicals to hydrogen
radicals and water. Hydrogen radicals are not as effective at initiating the chain
branching reaction sequence as hydroxy radicals are. In the toluene oxidation mechanism
proposed by Emdee et al. [21], which is a reference for the LLNL detailed toluene
mechanism [22], most of the initiation and propagation reactions involving the hydroxy
radical have an analogous reaction with the hydrogen radical with similar rate constants.
This indicates that the conversion of hydroxy radicals into hydrogen radicals by adding
hydrogen rich reformate will not significantly affect the rate of the autoignition reactions
for toluene. Since the hydrogen in the fuel reformate works to slow down the
autoignition reactions in alkanes, reformate becomes more effective for TRFs as nheptane is added to decrease RON.
3.2.5 Effects of Compression Ratio
All eleven operating conditions were repeated for each of three compression ratios; 9.8:1,
11.6:1, and 13.4:1. An important effect of increasing compression ratio is an
improvement in the thermodynamic efficiency of the engine, resulting in higher torque
output for equivalent manifold pressures. Figure 3.48 shows how NIMEP at MBT spark
timing and stoichiometric, WOT conditions changes with rc. There is some scatter in the
points due to changing environmental factors such as atmospheric pressure, temperature,
and humidity. The data for each compression ratio is averaged and plotted as the
percentage increase of NIMEP with rc in Figure 3.49. Extrapolating the curve, the
maximum NIMEP increase is about 9% and occurs at an rc of about 14:1. The equation
for the quadratic curve fit is:
NIMEP
NIMEPMBT,9.8:1
0.126+
O.MBT 0.137RC - 0.00487RC 2
(3.2)
Figure 3.50 and Figure 3.51 show how NIMEP and the combustion retard required to just
avoid knock change with compression ratio for PRFs and TRFs, respectively. The data
shown is for stoichiometric, unboosted, WOT conditions. Plots for the other operating
conditions are available in Appendix E. As the compression ratio is increased, peak
cylinder pressures and temperatures increase, so the spark must be retarded to avoid
knock, resulting in increased combustion retard. Two effects influence NIMEP when rc
is increased, as illustrated in Figure 3.52. The first is an increase in torque due to
increased engine efficiency. The second is an increase in torque loss due to increased
75
combustion retard. At low compression ratios and mild amounts of spark retard, NIMEP
increases with rc. At higher compression ratios and higher amounts of spark retard,
NIMEP decreases with rc. For each fuel type and operating condition there is an
optimum r, where engine torque output is at a maximum. For the conditions shown in
Figure 3.50 and Figure 3.51 the optimum rc is about 11:1 for 95 RON fuels, which is
similar to compression ratios used in modem automotive engines. The tradeoff between
torque increase from higher thermal efficiency and torque decrease from retarding spark
to avoid knock is discussed further in Section 5.1.1 .
The average increase with compression ratio of combustion retard required to avoid
knock is shown in Figure 3.53, separated by fuel type and rc interval. Points with very
high combustion retard - greater than 300 - or with spark timing earlier than MBT were
omitted. The bars, which reflect one standard deviation in each direction, indicate that
there is significant scatter in the data. There do not appear to be any clear trends to the
scatter save that for X, which is discussed in Section 3.2.2 . The average increase of
combustion retard with compression ratio, about 3' CA per unit r", does not seem to be
significantly affected by fuel type or rc range.
76
11
A
A
A
A1
A
A
0
-
10.5
-
10
-
9.5
* Lambda = 1.0, Toluene
A Lambda = 1.3, Toluene
* Lambda = 1.6, Toluene
o Lambda = 1.0, PRF120
A Lambda = 1.3, PRF120
-9
c
-8.5
IL
* Lambda = 1.6, PRF120
0 0
A AD1
1
A
A
30
A
-8
13
A
-7.5
-7
-
6.5
-6
50
60
-10
0
10
20
30
40
Spark Timing (*BTC)
Figure 3.18 - Change of NIMEP with spark timing for a range of X; r, = 9.8:1, 1500
rpm.
1.05
mLambda = 1.0, Toluene
A Lambda = 1.3, Toluene
I
* Lambda = 1.6, Toluene
*A
t-
0.95 -
A
DA
I
I
a.
A
I
0.9
* Lambda = 1.0, PRF120
A Lambda = 1.3, PRF120
* Lambda = 1.6, PRF120
96
A
IW
A
I
z 0.85
0.8 -
A
I
96
I
0.75 0.7
A
-
-1 )
0
Spark Retard
10
(esp-OSP,MBT,
20
30
*ATC)
Figure 3.19 - Normalized decrease of NIMEP with spark retard for a range of X; r, =
9.8:1, 1500 rpm, 10.1 bar NIMEPMBT.
77
1.05
mRc = 9.8:1, Lambda = 1.0
A Rc = 9.8:1, Lambda = 1.3
I
*
*
A
*
I
iE
0.95
-
0.9
-
0.85
-
0.8
-
0.75
-
++
9.8:1, Lambda = 1.6
11.6:1, Lambda = 1.0
11.6:1, Lambda = 1.3
11.6:1, Lambda = 1.6
o Rc = 13.4:1, Lambda = 1.0
A Rc = 13.4:1, Lambda = 1.3
o Rc = 13.4:1, Lambda = 1.6
0W
w
Rc =
Rc =
Rc =
Rc =
~
A"
0Z
A*
A*
0.7
0
-10
20
10
Spark Retard
(Osp-OP,MBT.
30
0
ATC)
Figure 3.20 - Normalized decrease of NIMEP with spark retard for a range of X and
re; toluene fuel, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at
unboosted WOT X = 1.0.
50
* Lambda = 1.0, Toluene
A Lambda = 1.3, Toluene
* Lambda = 1.6, Toluene
4
1
40
I-
A
0
A+'
W
30
-
20
-
c0
o Lambda = 1.0, PRF120
* Lambda = 1.3, PRF120
* Lambda = 1.6, PRF120
IIt
0
100
0-
.0
-10
-10
0
10
Spark Retard
20
30
40
50
0
(OP- S,MBT, *ATC)
Figure 3.21 - Change of combustion retard with spark retard for a range of X; r, =
9.8:1, 1500 rpm, 10.1 bar NIMEPMBT.
78
1.05
*rc9.8,
I
,
0.95 -
+
(-
LU
0.9 -
i
a. 0.85 LU
Z
0.8
K >%.
0.75 -
+
4
7
.
0.7 - _
-1 0
Spark Retard
30
20
10
0
(0sp-Osp,MBTO 0ATC)
11.0, n1O.1
A rc9.8, 11.3, n10.1
* rc9.8, 11.6, nl0.1
Drc9.8, 11.0, n1O.1, prl120
A rc9.8, 11.3, nI0.1, prf20
* rc9.8, 11.6, n10.1, pr1120
X rc9.8, 11.3, n10.1, 15%ref
* rc9.8, 11.3, n10.1, 30%ref
* rc9.8,11.3, nl 1.6
+ rc9.8, 11.3, n13.1
-rc9.8, 11.0, n12.1
-rc9.8, 11.0, n14.1
rc9.8, 11.0, n14.1, 15%ref
3 rcg.8, 11.0, n14.1, 30%ref
A rc1.6, 11.0, n10.7
X rc1.6, 11.3, n10.7
* rc1 1.6, 11.6, n 10.7
* rcl 1.6, 11.3, n1 0.7, 15%ref
+ rcl1.6, 11.3, n10.7, 30%ref
-rcl1.6, 11.3, n12.2
- rcl1.6, 11.3, n13.8
*rc1.6, 11.0, n12.7
* rcl 1.6, 11.0, n14.7
A rcl1.6, 11.0, n14.7, 15%ref
X rcl1.6, 11.0, n14.7, 30%ref
* rcl3.4, 11.0, nl1.0
+ rcl3.4,11.3, n11.0
+ rc1 3.4, 11.6, n 11.0
11.3, n11.0, 15%ref
-1rc13.4,
-rc13.4, 11.3, n11.0, 30%ref
+rcl3.4, 11.3, n12.5
* rcl 3.4, 11.3, n14.0
* rcl 3.4, 11.0, n 13.3
Xrcl 3.4, 11.0, n1 5.4
* rcl3.4, 11.0, n15.4, 15%ref
*rcl3.4, 11.0, n15.4, 30%ref
Figure 3.22 - Change of normalized NIMEP with spark retard for a wide range of
operating conditions and compression ratios; toluene fuel except where noted, 1500
rpm.
1.05-
-Correlation
Y = 1 - 0.168((1 + 4.443E-3XA2)A0.5
RMS Error = 0.005
1-
-
LU
0.95
-
0.9
-
0.85
-
0.8
-
0.75
-
0.7-10
-
*
1)
20
10
Combustion Retard
0
0
( 50%- 60%,MBT,
30
0
ATC)
rc9.8, 11.3, n10.1
rc9.8, 11.6, n10.1
rc9.8, 11.0, n0.1, prfl120
A rc9.8, 11.3, n0.1, prf120
* rc9.8, 11.6, n10.1, p1f120
X rc9.8, 11.3, n1O.1, 15%ref
K rcg.8, 11.3, n10.1, 30%ref
* rc9.8,11.3, nl1.6
+ rc9.8, 11.3, n13.1
- rc9.8, 11.0, n12.1
rc9.8, 11.0, n14.1
rc9.8, 11.0, n14.1, 15%ref
o rc9.8, 11.0, n14.1, 30%ref
A rc1.6, 11.0, n10.7
X rc1.6, 11.3, n10.7
)K rc1 1.6,11.6, n 10.7
* rc11.6,11.3, n10.7, 15%ref
+ rcl1.6, 11.3, nl0.7, 30%ref
- rc1.6,11.3, n12.2
rc1.6,11.3, n13.8
Srcl1.6, 11.0, n12.7
* rc1.6, 11.0, n14.7
A rc11.6, 11.0, n14.7, 15%ref
X rc11.6, 11.0, n14.7, 30%ref
)K rc1 3.4, 11.0, n1 1.0
* rc13.4,11.3, n11.0
+ rcl3.4,11.6, n11.0
- rc13.4, 11.3, nl1.0, 15%ref
rc3.4, 11.3, n11.0, 30%ref
+ rc13.4, 11.3, n12.5
* rcl3.4, 11.3, n14.0
A rc13.4, 11.0, n13.3
x rc13.4, 11.0, n15.4
K rc13.4, 11.0, n15.4, 15%ref
o rc13.4, 1.0, n154 30%ref
*
o
++
0
rc9.8, 11.0, n1O.1
A
40
Figure 3.23 - Change of normalized NIMEP with combustion retard for a wide
range of operating conditions and compression ratios; toluene fuel except where
noted, 1500 rpm.
79
0.05
toluene
O~p 28--28*BTC
Se,=
0,,=16*BTC
90.04-
.-...... TRF near knock
12*BTC
-08,=
0 =8*BTC
--
0.03
-o
4.1P
0,,=4*BTC
0.02-
0.01 -
-30
-15
-15
0
0
Cran
60
45
30
is
75
Crank Angle (*ATC)
Figure 3.24 - Averaged heat release rate profiles for several spark timings for
toluene and for TRFs with octane numbers that result in near knocking conditions;
r,= 9.8:1, X = 1.3, 1500 rpm, MAP = 1.23bar (for NIMEPMBT 10.1bar).
40
-0-Toluene
-+- TRF, near knock
*
30 2
C
*0
,-
-
0-
10-
0-
E
0
-10
'
-10
0
I
I
10
20
I
30
40
0
Spark Retard (Op.-0sp,MBT9 ATC)
Figure 3.25 - Change of combustion retard with spark retard for toluene and for
TRFs with octane numbers that result in near knocking conditions; rc = 9.8:1, x =
1.3, 1500 rpm, MAP = 1.23bar (for NIMEPMBT = 10.1bar).
80
-
rc =13.4:1
rc = 11.6:1
= 9.8:1
-- - .rc
'02 5-
20
15 ....
10
5
E
20
-5 .....
80
1
.90
Lambda
1.4
100
::.
::.
1610Fuel
RON
Figure 3.26 - Effects of X and PRF fuel RON on combustion retard to just avoid
knock; 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
-
-- ..
rc = 13.4:1
rc = 11.6:1
rc = 9.8:1
-
30
25
-
-...-
20
10
-..
..
5
--
-5
100
1.4
Lambda
80
. --
1
1.6
110
Fuel RON
Figure 3.27 - Effects of X and TRF fuel RON on combustion retard to
just
avoid
knock; 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
81
35
0
Torque Loss
30
- 2UO
25
-
- +- -PRF95
. --
Ii-.-
20
-
15
-
10
-
-.
E
0
TRF95
---
UTG96
. PRF90
-
.:
10%
- -0-
0.
(L)
- --
0'&
' ,-.---- ---
TRF90
UTG91
3%
5- 1%
0-
-5
i
0.9
1.1
1.3
Lambda
1.7
1.5
Figure 3.28 - Effect of A on combustion retard to just avoid knock for UTG fuels; r,
= 11.6:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
12
PRF95
---
-e- PRF90
8-
-e--PRF85
- PRF80
4-
-*-
0
E)
.0
-11
0 1
- - --
0
-0- TRF90
TRF85
-0-
U
TRF95
TRF80
-4-
.8
-
-12
I
1.2
1.4
1.6
Lambda
Figure 3.29 - Change of combustion retard to just avoid knock with increasing X; r,
= 9.8:1, 1500 rpm, MAP at X> 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
82
12 -
-M-
PRF105
-h- PRFIOO
8-
- PRF95
PRF90
-U-
4-
.0
-A- TRF100
01:
- . - %- A. ':--:
E
-0
.--
-4-
-0-
TRF95
-I-
TRF90
- UTG96
-
-
-
-a -
-11
-0
UTG91
-8
-12
1
1.2
1.6
1.4
Lambda
Figure 3.30 - Change of combustion retard to just avoid knock with increasing X; r,
= 11.6:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
-
12
8
-)K--
PRF110
-X-
PRF105
-+-
PRFI 00
...
-+-PRF95
-a
4
0
-
-
--
-
-12
TRF105
-A-
TRF100
-0-
TRF95
- -o- -UTG96
-4-
-8
-X-
-
'
I
1.2
1.4
1.6
Lambda
Figure 3.31 - Change of combustion retard to just avoid knock with increasing X; r,
= 13.4:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
83
35
I-
Torque Loss
-PRF115
-W-PRF110
30 -
-X-PRF105
25
l
-e-PRFI
00
CD
9
20 - 10%
-+PRF95
U
PRF90
150
7E 10 - 3%
5
.0
E
0
1%
-----
0 -------
0.
-5 i
I
I
1
9
10
11
12
---------1
13
1
15
1
14
15
16
NIMEP (bar)
Figure 3.32 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for PRFs; r,= 11.6:1, X = 1.0, 1500 rpm.
35
-
Torque Loss
--
30 - 20%
oI
TRF105
-*--TRF100
0
IP
25
9
--
TRF95
-
-U--TRF90
20 - 10 0
15
-
10
- 3%
5
1%
0
0
e-0000
U
0 - -
-------------
-5
9
10
11
12
13
114
15
16
NIMEP (bar)
Figure 3.33 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for TRFs; r,= 11.6:1, X = 1.0, 1500 rpm.
84
16
X TRF105
A TRF100
15 -
IL
.0
14
-
13
-
12
* TRF95
* TRF90
-WOT
torque loss
-
20% boost
..
.......
..................................
-40%
boost
-
11 10a
5
0
-5
15
10
Combustion Retard
25
20
(O05%-00%MBT,
30
35
*ATC)
Figure 3.34 - Lines of constant MAP calculated from Eq. (3.1) imposed on data
from Figure 3.33. Horizontal lines are drawn at NIMEPMBT- Vertical distance
between curves and lines is torque loss from retarding spark to avoid knock.
..
35 .
-.........
--........
Rc=13.4:
Rc=1 1.6:1
Rc=9.8:1
.
35
25.
Boosted
....
8
20
(b-
0
- .....:
1 5 ......
*0
L)
10
80.
...
-...
...--..
~
.............
...
90
100
RON
....
::.-..
16
.1n12.............
110
120
8
10
Boosted NIMEP (bar)
Figure 3.35 - Effects of boosted NIMEP and PRF fuel RON on combustion retard to
just avoid knock; X = 1.0, 1500 rpm.
85
-Rc=1 3.4:1
Rc=1 1.6:1
35 . ...........3 0
Rc=9.8:1
--...---....-
--
-........
--
.0 20
g15 .
-..........
10
E
0
........-....
2..........
1 1....
R8 N
120
t......
.N ME (
r
0
(bar)fe ONo omutonrtr
Figure 3.36 - Effects of b~~~oosted NIMEP
00
to just avoid knock; X = 1.0, 1500 rpm.
12028
%IME,(bar
sPRF95
35
oos1e
030 - 20%.
.TR9
-+-UTG96
25 - Torqu-Los..-4-P-F9
.a
15--
C
1 30
03%
--
0
-
--------
--
--
5 - 1%
E
.5 -I
9
10
1
1
11
12
13
NIMEP (bar)
Figure 3.37 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for UTG96, PRF95 and TRF95; re = 11.6:1, X = 1.0, 1500
rpm.
86
15
0
a)
10
-
0
0
**
D
E
0
*
PRFs
o TRFs
S
S
Se
0
UTG96
0
0
I-
*
-
Linear (PRFs)
-
Linear (TRFs)
-
Linear (UTG96)
0
S00
5
0
-OOC
b 0
0
1.4
1.3
1.2
1.1
1
massfiowalrmassflowalr,unboosted
Figure 3.38 - Increase of combustion retard with increased airflow rate from
boosting at 1500 rpm, PRFs at X = 1.0 and X = 1.3, TRFs at X = 1.0 and X = 1.3, and
UTG96 at X = 1.3. Data is from all three compression ratios.
15
S
+ UTG96
S
0*
7E
*
e
PRFs
o TRFs
0
0
10
-
0
-
Linear (PRFs)
**
Linear (TRFs)
-
5
*
000-0
E
0
Se0*
S.
5-
59
*
0
*
o00
Q-6
Linear (UTG96)
0
e-<
5
P0
000
0
0
1
2
3
4
NIMEP-NIMEPunboosted (bar)
Figure 3.39 - Increase of combustion retard with increased NIMEP from boosting
at 1500 rpm, PRFs at , = 1.0 and X = 1.3, TRFs at X = 1.0 and k = 1.3, and UTG96 at
X = 1.3. Data is from all three compression ratios.
87
16
35
---
PRFI10
C.R.
U
30
-
25
-
20
-
15
... ;.X
.
.- - ---
14
I-
0
E
15 I10
13 I
uJ
- -- -- - -
---
- -
-
-
--
-
-X-
PRF105
C.R.
--
PRF100
C.R.
-+--
PRF95
C.R.
---
PRF90
C.R.
'PRF110
-
--
-
-x- -PRF105
NIMEP
12
5-
NIMEP
.0
E
04
11
0-
e-
-
-PRFIOO
NIMEP
- PRF95
- -
NIMEP
10
-5
0
5
15
10
20
30
25
-
-
35
PRF90
NIMEP
Reformate Fraction (%)
Figure 3.40 - Decrease of combustion retard to just avoid knock and associated
NIMEP increase with increased reformed fuel fraction for PRFs; r, = 11.6:1, x = 1.0,
1500 rpm, MAP for 40% boost (NIMEPMBT 1.4 times unboosted NIMEPMBT).
35
- - 16
-- 4-TRF105
C.R.
30
-- 15
.~~~~~~~.
25
. . .
. .
. -.
-
-. -. . .
--
TRF100
C.R.
- x
-
--
U
TRF95
C.R.
-14
20
-
15
-
0
E
10
-
-U-TRF90
-- 13 a.
------------
-- -
0
-x- TRFIO5
NIMEP
--- - -
-- 12
(..
E
C.R.
-
- -A-
TRF100
NIMEP
5:
-
TRF95
-
.a. TRF90
11
0-5
-
0
5
10
15
20
25
30
10
NIMEP
NIMEP
35
Reformate Fraction (%)
Figure 3.41 - Decrease of combustion retard to just avoid knock and associated
NIMEP increase with increased reformed fuel fraction for TRFs; re = 11.6:1, x = 1.0,
1500 rpm, MAP for 40% boost (NIMEPMBT 1.4 times unboosted NIMEPMBT).
88
16
* TRF95 C.R.
15
-40%
boost
.........
1,,,..................................................................
-
i
14
c-
30% ref.
15% ref.
-
torque loss
40% boost
13 -
U
0.
z
12 20% boost
11
-
10
-
WOT,
no boost
a
-5
0
5
30
25
20
15
10
35
0
0
Combustion Retard (6 o%- 60%MBT, *ATC)
Figure 3.42 - Increase of combustion retard with boost and decrease of combustion
retard with reformate fraction for TRF95; r, = 11.6:1, X = 1.0, 1500 rpm. Curve for
NIMEP vs. combustion retard at 40% boost calculated from Eq. ( 3.1).
12
* TRF90 C.R.
-
11
-
..
.................... ...
............................
torque loss
15% re .
IL
(U
Rc=13.4:1
10 -
11.6:1
z
9.8:1
9
13.4:1
8
-5
0
5
10
15
20
25
30
35
0
0
Combustion Retard (Oso%- 5O%MBT9 ATC)
Figure 3.43 - Increase of combustion retard with r, and decrease of combustion
retard with reformate fraction for TRF95; X = 1.3, 1500 rpm, MAP set to match
NIMEPMBT at unboosted WOT X = 1.0. Curve for NIMEP vs. combustion retard for
rc= 13.4:1 calculated from Eq. ( 3.1).
89
Torque Loss
)
30 - 20%
--
UTG96
C.R.
-E-
UTG91
C.R.
25
-
+- PRF95
-
TRF95
C.R.
PRF90
C.R.
' -C.R.
9 20
10%
215
W:
-
10 - 3%
----..
'-.
-C.R.
C
.0
E
U
04
-,--
--
----
TRF90
-
5 - 1%'
0 -- - - - - - - - - - - - - - - - - - - -'----50
10
5
15
20
35
30
25
Reformate Fraction (%)
Figure 3.44 - Decrease of combustion retard to just avoid knock with increased
reformed fuel fraction for UTG fuels with PRFs and TRFs for comparison; r, =
11.6:1, X = 1.0, 1500 rpm, MAP for 40% boost (NIMEPMBT 1.4 times unboosted
NIMEPMBT).
----.
....
Rc= 13.4:1
Rc=11.6:1
Rc=9.8:1
40
030
G30 -
..........
-...
...
...
25
20
15
.10
E
0
-5
-.
..........
....
...
20
Reformate Fraction (%)
100
-...
010
30
110
RON
Figure 3.45 - Effects of reformed fuel fraction and PRF fuel RON on combustion
retard to just avoid knock; k = 1.0, 1500 rpm, MAP for 40% boost (NIMEPMBT 1.4
times unboosted NIMEPMBT).
90
..............
........
R c= 1 3 .4:1
Rc=1 1.6:1
......
Rc=9.8:1
.
3-~
20'R
5 ..
8 215-.8
10--
0
.5-...
0~2
........R...
Reformats Fraction (%)
3
2....
-......
5
Figure 3.46 - Effects of reformed fuel fraction and TRF fuel RON on combustion
retard to just avoid knock; X = 1.0, 1500 rpm, MAP for 40% boost (NIMEPMBT 1.4
TR.0
'..
...
80
times unboosted NIMEPMBT).
30
* PRFs
00
10
.......
-.....
a TRFBO-TRF95
-20
5 0
-
.2 15TRF1O5
-x
E10 -5
0
5
25
Reformate Fraction (%)
10
15
20
30
35
Figure 3.47 - Decrease of combustion retard with increased reformed fuel fraction
at 1500 rpm, PRFs at X = 1.0 (40% boost) and X = 1.3, TRFs at X =1.0 (40% boost)
and X=1.3, and UTGs at A = 1.3. Data is from all three compression ratios.
91
11.4
A 1st Toluene
A 2nd Toluene
11.2
-
* Ist PRF120
* 2nd PRF120
I..
.0
* 1st Toliso
* Average
11 10.8
-
10.6
-
10.4
-
-
Poly. (Average)
AA
2
10.2 10 -9
10
12
11
14
13
Figure 3.48 - Increase of unboosted NIMEP at MBT spark timing with r, for three
fuel types; X = 1.0, 1500 rpm. Closed symbols represent the first high load runs with
new pistons, open symbols represent runs made after several engine hours at high
load.
9
y = -4.87E-01x 2 + 1.37E+Olx - 8.74E+01
8
0
j7
6K
0
-5
w
4
3
2
0
9
10
12
11
13
14
Rc
Figure 3.49 - Normalized average increase of WOT NIMEP at MBT timing with r,;
) = 1.0, 1500 rpm. Raw data is from Figure 3.48.
92
35
11
~
0to-
30
e(
b ••
0
.. -
~PRF105
........ --~
C.R.
• • • • • • • b • • • • • • • • • • • • "JO.
~--.-
.:
25
III
~
""'-PRF100
C.R.
10
--'-PRF95
C.R.
~
0
• • • • • • • • • • e
on
cp 20
~
9
~
0
on
s
'0
15
---PRF90
C.R.
:E
• -x· ' PRF105
NIMEP
8 Z
(1)
..
~
ecu
Il.
W
~
cu
0::
-
10
t:
• -co- 'PRF100
0
U)
:::l
NIMEP
5
.c
7
E
0
0
-
-¢-
-
-0.
' PRF95
NIMEP
------------------
0
'PRF90
NIMEP
6
-5
9
11
10
12
13
14
Rc
Figure 3.52 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased rc for PRFs; A = 1.0, 1500 rpm, unboosted WOT.
40
0
to-
11
35
.
e(
0
b •••
.:III 30
_
.. -
.. _ ..
---'_
... _--
~
on
cp
.
--
~
25
9
~
0
on
S
..
cu
w
15
-
"JO..
- -<>-
:E
8 Z
(1)
t:
~
cu
Il.
0::
0
-
---TRF90
C.R.
e
20
'0
~
--+-TRF95
C.R.
10
~
0
---TRF100
C.R.
10
-
o[J.
TRF100
NIMEP
TRF95
NIMEP
TRF90
NIMEP
U)
:::l
.c
5
7
E
0
0
0
-5
+-----~------~------.-----~------+6
9
10
12
11
13
14
Rc
Figure 3.53 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased rc for TRFs; A = 1.0, 1500 rpm, unboosted WOT.
93
A TRF100
C.R.
11 -
............
torque loss
.... ........... .
10 -
* TRF95
C.R.
l
t
.................. ........
* TRF90
C.R.
-9.8:1
(L
-11.6:1
(U
9 -13.4:1
8
7
0
-5
5
10
15
Combustion Retard
20
0
35
30
25
( 60%-8o%MBT,
40
0
ATC)
Figure 3.52 - Curves for NIMEP vs. combustion retard from Eq. (3.1) imposed on
data from Figure 3.51. Vertical distance between curves and horizontal lines at
NIMEPMBT is torque loss from retarding spark to avoid knock.
0-%6
4
3
II
I I
0
(2
E
0
0
9.8:1-11.6:1
PRF
11.6:1-13.4:1
PRF
9.8:1-11.6:1
TRF
11.6:1-13.4:1
TRF
Figure 3.53 - Average increase in combustion retard to avoid knock over all
operating conditions considered, organized by fuel type and compression ratio
interval. Bars represent one standard deviation.
94
CHAPTER 4. CHEMISTRY MODELING
The results of a knock model under development at MIT were compared to experimental
data to ascertain the sensitivity of the model to input parameters and define the range of
the models validity. It was found that high accuracy in the determination of initial
conditions and thermodynamic environment is required for consistent results. Calibration
of baseline initial conditions was required to match experimental data. The model
successfully predicted the response of PRFs to compression ratio and air-fuel ratio and
the response of TRFs to boost. The difference between the response of PRFs and TRFs
to air-fuel ratio was also captured. Constant volume simulations were used to help
identify the mechanism by which hydrogen and carbon monoxide affect autoignition.
4.1 BACKGROUND
Mechanisms for predicting the response of autoignition time to thermodynamic
conditions are available in varying levels of complexity and accuracy. Most assume that
the rates of reaction of the end-gas have an exponential temperature dependence
combined with a pressure dependent term. One of the simplest forms is to invert a single
empirical reaction rate equation to define an induction time and then integrate the
induction time over the evolution of the pressure and temperature profile of the end-gas.
Although useful for first-order approximations, this method is unable to reproduce the
complicated temperature and pressure dependence of hydrocarbon autoignition reaction
mechanisms. To better replicate the chemical kinetics of hydrocarbon oxidation,
mechanisms involving multiple reactions have been developed. These range from simple
or "reduced" mechanisms with several species and reactions to "detailed" mechanisms
with thousands of species and reactions. One can still consider detailed mechanism to be
reduced from millions of possible species and reactions.
This study utilized detailed mechanisms designed by Lawrence Livermore National
Laboratories. The LLNL PRF mechanism is a combination of the LLNL isooctane
mechanism [23] and n-heptane mechanism [24]. It was selected because it has been
shown to be valid over a wide range of thermodynamic conditions and because it is well
supported by LLNL researchers. The TRF mechanism was created for this study by
LLNL researcher William Pitz. It is a combination of the n-heptane mechanism and a
toluene mechanism [22].
95
4.2 MODEL DESCRIPTION
The model is designed to simulate an adiabatic homogenous element of end-gas subjected
to the cylinder pressure generated by compression and combustion. It was assumed that,
prior to autoignition, pressure is uniform throughout the cylinder. The simulation starts
at the time of intake valve close and proceeds until autoignition or until a specified point
late in the expansion stroke if autoignition does not occur. The time of autoignition is
assumed to be the point at which the temperature rises faster than a critical rate. A rate of
2 x 106 K/s, or 222 K/0 CA at 1500 rpm, was found to give good accuracy while
maintaining numerical stability.
4.3 GOVERNING EQUATIONS
The structure of the simulation is shown in Figure 4.1. The rate constants and species
thermodynamics are taken from the PRF or TRF detailed reaction mechanisms. The
initial conditions consist of the molar concentrations of the reactants and the initial
temperature. The thermodynamic constraint for the engine knock simulation is the
cylinder pressure. For the constant volume simulation it is the volume, which is
calculated from the initial pressure using the ideal gas law.
The molar production rates for each species in the mechanism are calculated according to
the equation:
dC,
dt
k
U
' j
...
(4.1)
rx
Where C, is the molar concentration of species i, Jis the set of reactions for which i is a
product (both forward and reverse), Cs are the concentrations of the reactants in
reactionj, and ovj is the stoichiometric coefficient for species i in reactionj. The rate
coefficient, k is generally calculated according to the Arrhenius equation:
ki = A T"' exp
~ I
RT)
(4.2)
Where T is the temperature, R is the ideal gas constant (molar), and A1 , nj, and Ej are,
respectively, the pre-exponential factor, the temperature exponent, and the activation
96
energy for reactionj given by the reaction mechanism. The rate constant is multiplied by
a pressure dependent term for some reactions. The internal energy and specific heat for
each species is calculated from a polynomial fit to temperature given by the
thermodynamics section of the reaction mechanism.
)
-'N
Equations
molar production rates from reactions
conservation of mass
equation of state (Ideal Gas Law)
conservation of energy
Inputs
species & species thermodynamics
reactions & reaction rate constants (A,E,n)
initial conditions
const raints (ie. P(t) or V(t))
Numerical
Integrator
Outputs
species concentrations vs. time
thermodynamic state (ie. temperature) vs. time
autoignition time (derived)
Figure 4.1 - Structure of an autoignition simulation.
For the engine knock simulation using cylinder pressure as a constraint the ideal gas law
is represented by:
p = MaveP
RT
(4.3)
Where p is the density, Mave is the average molecular weight for the mixture, and P is the
total pressure. The equation for conservation of mass is written for each species, and
takes the form:
dc
dt
dC.
dt irX
~ 1 dP
(P dt
1 dT 1
T dt
(4.4)
97
Where the first term is the change of concentration due molar production from reactions
and the second term is the change of concentration due to changing mixture density. The
base equation for conservation of energy is:
dU = SQ - SW
(4.5)
Since heat transfer is neglected, 6Q is zero. Internal energy, dU changes due to the
change of thermodynamic state and due to changes in mixture composition:
dU = cVdT + I dmiu,
(4.6)
Where c, is the specific heat at constant volume, dmi is the differential change of mass of
species i, and ui is the internal energy for species i. Using the definition of work and the
ideal gas law:
SW = Pdv = RdT - vdP
(4.7)
Where v is the specific volume of the mixture. Equating the two and collecting terms:
cdT +Z dmju, =vdP
(4.8)
Where c, is the specific heat at constant pressure. For use in the model molar quantities
were used and the equation was rearranged and differentiated in time:
,Pp dT
Mave dt
~ dC
'idt
dP
dt
For the constant volume simulations the calculation of reaction rates and species
thermodynamics is the same as described above. The ideal gas equation of state and
conservation equations are similar to those shown above, but rearranged to use volume as
the constraint instead of pressure.
4.4 ENGINE KNOCK SIMULATIONS
The Jacobian IDE solver by Numerica Technology LLC was used to integrate the engine
knock simulation. Numerica Technology's Open Chem Pro routines were used to
calculate molar production rates and species thermodynamics. The rest of the equations
98
were written directly into the IDE. Each simulation from intake valve closing time to
autoignition takes approximately 30 minutes to run on a modem PC.
4.4.1 Selection of Input Parameters
Besides the chemistry mechanism, the other key inputs to the knock simulation are the
pressure trace, the initial species concentrations, the initial pressure, and the initial
temperature. For this application these inputs were derived from the detailed
experimental data corresponding to the results presented in Section 3.2 .
Analysis of pressure data by Topinka et al. indicates that at borderline knocking
conditions, cycles with peak pressure locations that are more than one standard deviation
earlier than the mean are most likely to knock [14]. Thus, for previous work, data from a
cycle with a peak pressure that is one standard deviation earlier than the mean was
selected as an input to the knock model. The method used in this work is similar, but
needed to be modified for the conditions under which data was collected. Instead of
using peak pressure as an indicator of combustion phasing, the location of 50% mass
fraction burned was used. This was necessary because when spark is significantly
retarded, as it is for many of the data points in this study, there can be two pressure
maxima, one near TDC from compression, and a later one from combustion. Also, since
the data taken in this study was at a spark timing one degree later than that for borderline
knock, not as many cycles are likely to autoignite. Observations of pressure data, such as
that shown in Figure 4.2, indicate that under near-knock conditions approximately the
earliest 10% of cycles show mild autoignition. Thus the pressure trace selected from a
set of data for input to the knock model was one with a location of 50% mass fraction
burned that is earlier than 90% of the cycles in the set (i.e. 1 0 th percentile early).
There are three components to the intake charge that must be taken into account when
determining the in-cylinder mixture composition. They are intake air, fuel, and residual
gas. The air flow and fuel flow were taken directly from experimental measurements.
The humidity of the air was measured and taken into account. Residual mass fractions
were approximated using linear fits to results from a Ricardo WAVE 1-dimensional fluid
flow and engine simulation. An example of the simulation results is shown in Figure 4.3.
Since the valve overlap on the experimental engine is small, most of the residual is from
the cylinder clearance volume, which is at approximately atmospheric pressure at the end
of the exhaust stroke. Increased residual temperature decreases density and thus
decreases residual mass, which explains the downward trend of residual fraction with
99
increased spark retard. It also decreases with increased charge mass due to increased
dilution. For stoichiometric boosted operation the model residual output scales inversely
with charge mass to within 2% of the absolute value.
45 -
40
Earlier than 90% of cycles
Later than 90% of cycles
-
INi
30
U)
M.
-"D~C"-
4,,
-
*44
-
.
25 -
20
15
-
i
0
10
20
30
40
50
Crank Angle (*ATC)
Figure 4.2 - A set of 90 cycles of pressure data taken under near knocking
conditions with PRF90; 1500rpm, WOT, X = 1.0. The solid lines are cycles that have
a location of 50% mass fraction burned that is earlier than 90% of the cycles.
Initial temperature was estimated for unboosted data points using the ideal gas law with
the volume at intake valve close, the measured manifold pressure, and the number of
moles of charge in the cylinder calculated from air flow, fuel flow, and residual fraction:
Tec = PintVeIVC
(4.10)
nchR
Where Pin, is the intake manifold pressure and nch is the number of moles of charge in the
cylinder. Since compression begins before the actual intake valve closing time, an
100
effective intake valve closing time, VeIvc, was estimated by comparing cylinder pressure
data to a polytropic compression curve. Figure 4.4 shows that the two curves become
aligned at about 400 ABC.
5 -
-
4
0
-
UU
0L
32-
1-~
-- Rc=9.8:1
y(9.8) = -0.0354x + 4.1435
-- Rc=1 1.6:1
y(11.6) = -0.0381x + 3.9022
+Rc=1 3.4:1
y(13.4) = -0.0414x + 3.6633
0
10
0
20
30
e50%mfb
( 0ATC)
40
50
Figure 4.3 - Residual fractions calculated by a Ricardo WAVE engine simulation;
X = 1.3, 1500 rpm, MAP = 1.22 bar.
2C1.8 -
-
Measured pressure
---
Polytropic compression (k=1.29)
Effective,
IVC,'
-
CL
1.4
1.2
-
1
180
20 0
220
240
260
Crank Angle (*ATC)
Figure 4.4 - Comparison of measured cylinder pressure to polytropic compression
to find effective intake valve closing time; 1500 rpm, X = 1. 5.
101
Under boosted conditions the throttle had to be closed part way to control air flow. This
induced pressure fluctuations in the intake manifold and made experimental manifold
pressure measurements unreliable. For these cases, the WAVE engine model was used.
For the unboosted baseline case, WAVE predicted an initial temperature 11 K lower than
the experimental value calculated from the ideal gas law. The error is most likely due to
errors in modeled heat transfer coefficients and surface temperatures. Instead of using
the WAVE predicted values directly, the variation of the predicted initial temperature
was added to the experimentally estimated initial temperature from the baseline case.
The results are shown in Table 4.1. The intake manifold pressure was then estimated
using the ideal gas law:
P=
(4.11)
Ivc
VeIVC
nch
Table 4.1 -Temperatures at intake valve close estimated by WAVE engine model.
TeIVC
(K)
X=1.0,
unboosted
X = 1.3,
X = 1.6,
20%
40%
boosted to
boosted to
boosted
boosted
NIMEPMBT,
maintain
maintain
NIMEPMBT,
NIMEPMBT
NIMEPMBT
X = 1.0
=
1.0
WAVE
predicted:
339
339
339
336
334
Adjusted
Adjistd
351
(from ideal
351
351
348
346
experimental:
gas law)
I
4.4.2 Pressure-Temperature Profiles
In order to gain an understanding of the temperatures that SI engine end-gas is subjected
to in response to cylinder pressure, simulations were run with the chemical kinetics
disabled. The results help to explain the trends in knock behavior with changes in
operating conditions.
Figure 4.5 shows the pressure and temperature profiles for two air fuel ratios, one with a
X of 1.0 and one with a X of 1.6. Spark timing is set for MBT and NIMEP is the same for
both cases. Even though cylinder pressure is higher for the lean case, the peak end-gas
temperature is slightly lower. For adiabatic compression, the increase in temperature is
102
dependant on the pressure ratio through which the mixture is compressed. The increased
initial pressure from boosting to maintain constant NIMEP decreases the pressure ratio
and thus decreases the temperature ratio. The decreased pressure ratio from boosting
more than compensates for the increase in temperature from the increased y for lean
mixtures.
1000
100
1.0 Pressure
-Lambda
90 - ----- Lambda 1.6 Pressure
900
-Lambda 1.0 Temperature
- - - Lambda 1.6 Temperature
80
800
70 cc
60
700
50 cc
600
40 -
500
30 20 -
-
-400
10 -
0 1
-90
1
-75
1
-60
300
-45
-30
-15
0
15
30
45
Crank Angle (*ATC)
Figure 4.5 - Experimental pressure profiles and predicted non-reacting end-gas
temperature profiles for stoichiometric and lean air-fuel ratios; iso-octane fuel, 1500
rpm, r, = 9.8:1, NIMEP = 10.1 bar, MBT spark timing.
Pressure and temperature profiles for boosted operation are compared to unboosted
operation in Figure 4.6. As with lean operation, although the pressure for the boosted
case is increased, the pressure ratio remains relatively constant, so peak temperatures do
not change significantly.
Increasing compression ratio increases both pressure and temperature, as shown in Figure
4.7. Inlet pressure remains constant, but peak pressure increases. This increases the
pressure ratio and so increases the temperature ratio. The initial temperature is slightly
lower due to decreased residual fraction, but its effect is insignificant compared to the
increase from higher compression pressures.
103
100
90
-
1Obar NIMEP Pressure
..
l..
14bar NIMEP Pressure
-1Obar
80
I 1000
- - - 14bar NIMEP Temperature
- 800
70
IC
40
~0
.0
I-
900
-
NIMEP Temperature
60
-
700
2
-
600
CL
-
500
-
400
50
40
30 -
E
20 10 .-
300
0
-90
-75
-60
-45
-30
-15
0
15
30
45
Crank Angle (*ATC)
Figure 4.6 - Experimental pressure profiles and predicted non-reacting end-gas
temperature profiles for unboosted and boosted conditions; iso-octane fuel, 1500
rpm, r, = 9.8:1, X = 1.0, MBT spark timing.
100
1000
a
9 8:1 Drafssre
I
------ 13.4:1 Pressure
9.8:1 Temperature
90
80
I-
-900
- - 13.4:1 Temperature
70
cc
.01
60
-
U)
50
-
40
-
30
-
-
800
-
700 2
-e
I-
-600
E0.
20 10
-
400
-
300
-
-90
-75
-60
-45
-30
-15
0
15
30
45
Crank Angle (*ATC)
Figure 4.7 - Experimental pressure profiles and predicted non-reacting end-gas
temperature profiles for low and high compression ratios; iso-octane fuel, 1500 rpm,
MAP = 1 bar, MBT spark timing.
104
Figure 4.8 shows the differences in temperature profiles between mixtures of air and
PRF 100 (iso-octane) and mixtures of air and TRF 100 (82% toluene). Toluene has lower
specific heats than iso-octane does, which causes it to have a higher y. At 600 K the
stoichiometric PRF mixture has a y of 1.311 while the stoichiometric TRF mixture has a y
of 1.323. For adiabatic compression a higher y increases the change in temperature in
response to a given change in pressure. The result is an increase in peak temperature.
I 1000
100
90 -
Pressure
PRF100 Temperature
-
- 900
- - - TRFIOO Temperature
80 70
VU
IL
-800
I-
-
60
I-
- 700
2
- 600
E
/-
50
40
30
-
500
-
400
20
101
300
-90
-75
-60
-45
-30
-15
0
15
30
45
Crank Angle (*ATC)
Figure 4.8 -Predicted non-reacting end-gas temperature profiles with the same
pressure profile for PRF100 and TRF100 fuels; 1500 rpm, re = 9.8:1, MAP = 1.0 bar,
X = 1.0, MBT spark timing.
Since TRF fuels have a higher y, they are more sensitive to changes in pressure ratio than
PRF fuels are. The differences in y between TRFs and PRFs help to explain the
differences between the trends of knock limited spark timing with air-fuel ratio presented
in Section 3.2.2 . In the experiments the air-fuel ratio was adjusted by increasing the inlet
pressure. This reduced the ratio between inlet pressure and peak pressure. Since TRF
fuels have a higher y, the peak temperatures for TRF fuels decrease more with decreased
pressure ratio than they do for PRF fuels. The decreased peak temperature for TRF fuels
compensates for the increased reaction rates from increased absolute pressure. For PRF
105
fuels the decrease in peak temperature is not enough to compensate for the increased
reaction rates at elevated pressures, so more spark retard is required.
Figure 4.9 shows the end-gas temperature profiles for two PRF95 simulations, one with
the chemical kinetics enabled and one with the chemical kinetics disabled. For the
simulation with kinetics enabled the rapid temperature rise at 270 ATC represents
autoignition. The temperature from the reacting simulation does not depart from the nonreacting one until about 100 CA before autoignition at a temperature of about 950 K.
60
50
-
I 1200
Pressure
Inert Temperature
- - - Reacting Temperature
- 1100
- 1000
40
-
-
I.- L 30 -
900
- 800 L
0
- 700
20 -
E
- 600
10 -
0
-75
-60
-45
-30
-15
0
15
30
-
600
'
400
45
Crank Angle (*ATC)
Figure 4.9 - Predicted end-gas temperature profiles with the same pressure profile
for simulations with reaction kinetics disabled and enabled for PRF95; 1500 rpm, re
= 9.8:1, MAP = lbar, X = 1.0, O,p = 6'BTC.
4.4.3 Model Sensitivity
Qualitative observations of pressure data such as that shown in Figure 4.2 indicate that
under near knock conditions, autoignition occurs near where peak pressure would be in a
non-knocking cycle. To assess the accuracy required of the model input parameters, the
variation of the autoignition time with respect to the time of peak pressure for variations
of each parameter was noted. Figure 4.10 shows the effect of changing initial
temperature on predicted autoignition time for a simulation with PRF95. For autoignition
106
to occur near peak pressure, an initial temperature of about 423 K is required. This
temperature is about 70 K higher that that estimated using the method described in
Section 4.4.1 . Since the actual value of the initial temperature of the portion of the endgas that autoignites is not well known, initial temperature was used as a variable for
calibrating the model. Variation of autoignition time with initial temperature for
autoignition times near peak pressure is about 20 CA per 5 K. The rest of the results of
the sensitivity study are listed in Table 4.2.
50 -1200
45
-
Pressure
-
Temperature
",j=4iK
/
Tjng=41OK
1100
TjT=405K
40/
1000 *
35 -
E
Tinj=400K
1000
30
900
25
800
20
0
30
15
Crank Angle (0ATC)
45
Figure 4.10 - Effect of initial temperature on predicted autoignition time in
increments of 5 K for PRF95; 1500 rpm, r, = 9.8:1, MAP = lbar, X = 1.0, ,
=
6'
BTC.
4.4.4 Comparison to Experimental Results
In order to assess the model's accuracy, it was compared to experimental results for
lambda sweeps, boost sweeps, and compression ratio sweeps for PRF95 and TRF95. For
each fuel, the initial temperature was calibrated at the baseline condition (rc = 9.8:1, X =
1.0, MAP = 1.Obar, spark timing for near knock) so that autoignition occurred at peak
pressure. As operating conditions were changed from the baseline conditions, the initial
107
temperature in the model was adjusted by the same amount that the estimated
experimental initial temperature changed.
Table 4.2 - Sensitivity of autoignition time to model input parameters. AOign.pp is the
distance in crank angle degrees from autoignition to peak-pressure.
Parameter:
Effect on
AOign-pp
(0CA):
Initial
temperature
Fuel octane
2' advance per
5 K increase
20 retard per 5
ON increase
number
Manifold
pressure (for
pegging P,,,)
Pressure trace
location of
50%mfb
20 retard per
20 retard per 30
retardof
i.ncr
Imicrease
____m_
For each operating condition, the model was run with pressure traces extracted from
spark timing sweep data under the same condition. The pressure traces were selected
from each set of data according to the criteria described in Section 4.4.1 . The spark
timing for which autoignition occurred closest to peak pressure was interpreted as the
knock limited spark timing, as shown in Figure 4.11.
Figure 4.12 to Figure 4.17 show comparisons of the model output to experimental data
along with the estimated experimental initial temperatures and those used for the
simulation. The combustion retard shown is that corresponding to the predicted knock
limited spark timing. The PRF simulations match the experimental results very well for
changes in compression ratio and changes in air-fuel ratio. The model under predicts the
combustion retard required to avoid knock for boosted operation. This is probably
mostly due to errors in the estimated initial conditions. Combined errors in initial
temperature and manifold pressure can add up to several degrees of error in the
simulation results. The chemistry mechanism, which has not been completely verified
under the conditions being explored, may also contribute to the difference. The TRF
simulations overestimate the effects of compression ratio and air-fuel ratio, but the
direction of the trends are correct. The errors may be partially due to errors in the initial
conditions, but this combined reaction mechanism has not yet been well tested and is
likely responsible for much of the difference. The boosted experiments appear well
matched.
108
10
1200
55 55
0
6 0BTC 4 BTC
20 TC
0
ep=8 BTC
50 -
1100
45
-
1000
IL.
35 -
E
a'
O*BTC
30 -
900
25-
800
20
-15
0
45
30
15
Crank Angle (*ATC)
Figure 4.11 - Pressure traces from several spark timings and corresponding
simulated end-gas temperature profiles; 1500 rpm, r, = 11.6:1, X = 1.0, MAP =
1.Obar. 40 BTC is interpreted as the predicted knock limited spark timing.
35
-
I-
30
-
25
-
20
-
0
+ PRF95
*
Model
PRF95 Exp.
Teivc,.st=343K
TKvc,.e=342K
T.jvc,.t=351 K
Ti.
15 -
=I4
Tingt=423K
10
-
50
E
0
0-5
9
10
11
12
13
14
Rc
Figure 4.12 - Comparison of model predicted combustion retard to experimental
combustion retard with increasing compression ratio for PRF95; 1500 rpm, X = 1.0,
MAP = 1.0bar.
109
35
- -- - od --lTjlf=30
TRF95 Model
---
- 0
Ti,,I=350K
30 - -- +-TRF95
Exp.
0
25
-
Tiit=348K
CP
20
-
15
-
,
T.ves,.t=344K
9
T..vc,.=342K
Tifft=357K ,10 T.jvc,..t=351K
.0
E
0
C.
5 0 -
'
1
9
10
-5
,I
12
14
13
12
11
Rc
Figure 4.13 - Comparison of model predicted combustion retard to experimental
combustion retard with increasing compression ratio for TRF95; 1500 rpm, X = 1.0,
MAP = 1.0 bar.
-
35
I-
30
-
- *
PRF95 Model
---
PRF95 Exp.
25 o
0
C
20
-
15
-
10
-
T.vc,..t=351 K
Ti"t=423K
T.vc,..t=351 K
T.vc,..t=3 5 1K
T101t=423K
T10ft=423K
0
5E
0
L)
01---- - - - -------------------------5 0.9
1.1
1.3
1.5
1.7
Lambda
Figure 4.14 - Comparison of model predicted combustion retard to experimental
combustion retard with increasing air-fuel ratio for PRF95; 1500 rpm, r, = 9.8:1,
NIMEPMBT= 10.1 bar.
110
35
-TRF95 Model
-
I-
30
-- +--TRF95 Exp.
0
0
So
0
(U
VSo
25
20
15
T.ivce..t=35 1 K
c-
a)
T.vc,..=351 K
10
Tinit=357K '
0
.0
E
0
0.
,
T.s 6.-t=351 K
5
Ti =357K
fT
0
, -- T=357K
Tj,
-5
1.5
1.3
1.1
0.9
1.7
Lambda
Figure 4.15 - Comparison of model predicted combustion retard to experimental
combustion retard with increasing air-fuel ratio for TRF95; 1500 rpm, r, = 9.8:1,
NIMEPMBT = 10.1 bar.
35
30
-
- *
PRF95 Model
-- *
PRF95 Exp.
TervC,..t=346K
TeivC,..j=348K
25 -
Teive,...=334
C
20
-
T~jvcA=:.K
Tikt=418K
15 -
,0-
TifTn=420K
C
"0
0
.0
E
.)
T,,t=423K
10 5 -
0-5
8
9
10
11
12
13
14
NIMEP (bar)
Figure 4.16 - Comparison of model predicted combustion retard to experimental
combustion retard with increasing boosted NIMEP for PRF95; 1500 rpm, r, = 9.8:1,
x = 1.0.
111
35
0
30
+TRF95
Model
-- +-TRF95 Exp.
25 T. 1vC,.m=346K
2
CP 20Tev,..t=348K
0
15-
T.vC,..=351 K
10
-
C
05
-
,
Tjflt=352K
'"t=354K
+T
Ti.K=357K
5-
E
0
*
----------------------
0-------5 -
8
9
I
I
I
I
10
11
12
13
14
NIMEP (bar)
Figure 4.17 - Comparison of model predicted combustion retard to experimental
combustion retard with increaing boosted NIMEP, PRF95; 1500 rpm, r, = 9.8:1, X
=
1.0.
4.5 CONSTANT VOLUME SIMULATIONS
Constant volume autoignition simulations were used to gain insight into the mechanisms
by which hydrogen and carbon monoxide affect autoignition of alkane fuels. Chemkin
v3.7 Aurora by Reaction Design was used as the solver. A constant volume perfectly
mixed adiabatic reactor simulator is included in the Aurora package. The user must
supply the reaction mechanism, species thermodynamics, and initial conditions. To
simulate conditions in an engine cylinder an initial pressure of 45 atm was selected. An
initial temperature of 875 K was chosen so that both the low temperature and high
temperature oxidation mechanisms were active and could be observed. For this study the
species mole fractions were chosen to represent a mixture of air and iso-octane with X
1.5. Hydrogen and carbon monoxide were added as 5.5% and 7% of the fuel energy,
respectively, similar to a mixture with 15% fuel reformed fraction.
112
=
4.5.1 Effects of Hydrogen Addition
Figure 4.18 shows a comparison between the temperature and OH radical concentration
profiles of autoignition simulations for iso-octane with air, and for iso-octane plus 5.5%
H2 by energy with air. The two-stage nature of autoignition is evident in the evolution of
the OH concentration. The first stage, consisting of rapid initiation and moderate heat
release, lasts for about 2 ms. The second stage lasts until just before the time of
autoignition. It is slower and less exothermic than the first stage. As expected, the
simulation predicts that the autoignition time for the case with hydrogen addition is
longer than that for just iso-octane.
1.E-02
3000
2500
- Temperature
Mole Fraction
-OH
-
CeH 18
C8H18
--
1.E-04
+ H2
0
I.E-06
2000 -
U.
L
0
E 1500
- 1.E-08
-
0
1000
0UU
----
-
.
1
0.000
0.003
0.006
00
0.009
I.E-10
1
1.E-12
0.012
Time (s)
Figure 4.18 - Temperature and OH concentration for constant-volume autoignition
simulations of iso-octane and of iso-octane with 5.5% H2 by energy; Tiit= 875 K,
Pi1it= 45 bar, X = 1.5.
To evaluate the effect of thermal dilution from the added hydrogen, a new species was
added to the simulation. The new species, nrH2, is thermodynamically identical to
hydrogen, but it is non-reactive. An autoignition simulation for which H2 was replaced
with nrH2 is compared to the original H2 addition case in Figure 4.19. The increased
thermal dilution from adding hydrogen accounts for about half of the overall effect. The
rest of the effect must be accounted for with changes in the chemical kinetics.
113
3000
1.E-02
-Temperature
Mole Fraction
-OH
1.E-04
2500 C8H18
+ nrH2
C8H,8
+H
2
0
I
2000 -
1.E-06 I
EU
E 1500 -
0.0
.1.E-08
1.E-10
1000 -
Soo
0.000
I.E-12
0.003
0.006
0.009
0.012
Time (s)
Figure 4.19 - Temperature and OH concentration for constant-volume autoignition
simulations of iso-octane with 5.5% H2 by energy, and of iso-octane with simulated
non-reactive H2 (nrH2); Ti it= 875 K, Piit= 45 bar, X = 1.5.
A limiting factor to the progression of the reactions leading to autoignition is the
availability of OH radicals to react with the alkane to produce an alkyl radical. Figure
4.20 shows the rates of reaction for OH radicals reacting with iso-octane and hydrogen
molecules, as compared to total OH consumption for the H2 addition and non-reacting
nrH2 addition simulations. For the reacting H2 case there is a slightly wider gap between
total OH consumption and OH consumption from the reaction with iso-octane.
Consumption of OH by H2 is negligible for the nrH2 case, but is approaching the same
order of magnitude as the reaction with iso-octane in the reacting H2 case. This indicates
that due to reactions with H2, less OH is available to initiate the chain branching reaction
sequence with iso-octane. A product of the reaction between H2 and OH are H radicals.
Although H radicals also react with fuel molecules to initiate the chain branching
sequence, Figure 4.21 shows that in both cases, the rate of this reaction is not significant
compared to the reaction with OH. Closer inspection of the reaction mechanism shows
that the rates for reactions of H with iso-octane are one to two orders of magnitude less
than those for OH.
114
1.E+02
+ nrH2
Sn 1.E+00
E
Total OH consumption
C811
+ H2
-
-C8H8+OH
H2+OH => H+H20
-
I-
=> C8H17+H20
0
Ii
0.
E
I .E-040
c.
1.E-08
0.009
0.006
0.003
0
0.012
Time (s)
Figure 4.20 - OH consumption by reactions with fuel molecules for constant-volume
autoignition simulations of iso-octane with 5.5% H2 by energy, and of iso-octane
with simulated non-reactive H 2 (nrH2); Tinit= 875 K, Pinit= 45 bar, X = 1.5.
1.E+02
-C8H8+OH
-
cv> 1.E+00E
CsH1 8
+ nrH2
-0
0
S1.E-02
=> C8H17+H20
C8H8+H => C8H17+H2
CBH 18
+ H2
-
0.
E
.
..
-
, ...
1.E-04
-
1.E-06
-
0
1.E-08
'I
0
0.003
0.006
0.009
0.012
Time (s)
Figure 4.21 - Rates of reaction of OH and H with iso-octane molecules for constantvolume autoignition simulations of iso-octane with 5.5% H 2 by energy, and of isooctane with simulated non-reactive H2 (nrH2); Tinit = 875 K, Piit= 45 bar, X = 1.5.
115
To test the theory that consumption of OH radicals by H2 is the reason for the nondilution portion of the increased autoignition delay, a different new species was added to
the simulation. The second new species, prH2 , is similar to the non-reacting species nrH 2 ,
except that it is allowed to react with OH only through the reaction "prH 2 + OH => H +
H2 0". The results of a simulation with prH2 are compared to a simulation with fully
reacting H2 in Figure 4.22. The autoignition times of the two cases are almost identical.
----
.
1.E-02
-
1.E-04
2000 -
-
1.E-06 t
E 1500 -
-
1.E-08
3000
-
-
Temperature
OH Mole Fraction
2500-
CBH 1 s C8H1 8
+ H2 +prH 2
0
ME
0
1000 -
500
0.0 00
-
.
.
.
.
0.003
0.006
0.009
0.012
I.E-10
1.E-1 2
Time (s)
Figure 4.22 - Temperature and OH concentration for constant-volume autoignition
simulations of iso-octane with 5.5% H2 by energy, and of iso-octane with simulated
H2 (prH2) that reacts only with OH; Tiit= 875 K, Pinit= 45 bar, ) = 1.5.
The proposed mechanism by which H2 slows down the reactions leading to autoignition
is shown in Figure 4.23. According the simulations run with the LLNL PRF mechanism,
H2 intercepts the OH radicals produced by the chain branching reactions and turns them
into H radicals. The H reaction with the alkane fuel is much slower than that for OH, so
the rate at which the chain branching reaction sequence is initiated is reduced. This
mechanism is similar to that proposed in [25], except that the participation of the H
radical in the reaction sequence has been neglected. An effect that is not included in the
reaction mechanism employed, and so was not investigated, is the reaction of H2 with
radicals that are intermediate to the chain branching reaction sequence. These types of
116
reactions, however, would have the same effect as the reaction of H2 with OH,
interrupting the chain branching sequence and slowing autoignition.
High Temperature
H202
-HO2
RH
Olefin -4
RH+O2
2
-H20
R.-
1-+--*
H
02.~
Low
Temperature
4
fonarkate
ROOHI
OH
IOr-
~OOROOH
-~*OROOH
0R0
Figure 4.23 - Proposed mechanism by which hydrogen impacts the autoignition of
alkane hydrocarbons. Base diagram is from Tanaka et al. [3].
4.5.2 Effects of Carbon Monoxide Addition
Before simulations were performed to investigate the effect of carbon monoxide addition,
the reaction rate for "CO + 02 => CO 2 + 0" was changed from that in the original LLNL
mechanism to that suggested by Scire et al. [26]. The original rate worked well for
simulations with normal concentrations of CO, but William Pitz at Lawrence Livermore
National Laboratories suggested the new rate for this application. A comparison of the
rates is shown in Figure 4.24.
Figure 4.25 shows a comparison between the temperature and OH radical concentration
profiles of autoignition simulations for iso-octane with air, and for iso-octane plus 7%
CO by energy with air. The simulation predicts that the autoignition time for the case
with carbon monoxide addition is longer by about half as much as it is for hydrogen
addition. This agrees qualitatively with the results of engine experiments that show that
addition of CO in quantities corresponding to that in plasmatron reformate has about onehalf to two-thirds the effect of H2 on knock [14].
117
8
- - Original Mechanism
--- Tsang & Hampson
Scire et al.
640S
CDi
%
\
\
'1%
%~
%%
2-
.2 0-2-4-6 40.0006
.
0.0010
0.0018
0.0014
1/T
Figure 4.24 - Comparison of the reaction rate constant for CO+02 => C02+O from
three sources. The rate from Scire et al. [26] was used in this study.
3000
-OH
2500
1.E-02
-Temperature
Mole Fraction
-
C8H18
H1
+8C
O
~ 1.E-04
+ CO
0
2000 -
- 1.E-06
LA.
- 1.E-08
E 1500 -
0
1000 -
500F0.000
- 1.E-10
1
0.003
1
1
0.006
0.009
I 1.E-12
0.012
Time (s)
Figure 4.25 - Temperature and OH concentration for constant-volume autoignition
simulations of iso-octane and of iso-octane with 7% CO by energy; Tiit = 875 K, Pinit
= 45 bar, X= 1.5.
118
As was done with hydrogen addition, a non-reactive species, nrCO, was created to
investigate the effect of thermal dilution. An autoignition simulation for which CO was
replaced with nrCO is compared to the original CO addition case in Figure 4.26. The
result shows that the increased thermal dilution from adding non-reacting carbon
monoxide extends the autoignition time just past that for reacting CO. This implies that
the main effect of CO is thermal dilution of the charge with a less reactive species, and
that the chemical kinetics that involve CO may actually work to slightly reduce
autoignition time. Although several reactions involving CO were found to influence
autoignition time, the changes made by adjusting those reactions were small compared to
the accuracy of the simulation, so no further conclusions were made.
3000
-
1.E-02
- --
Temperature
OH Mole Fraction
1.E-04
2500C8H16
+CO
COH 18
+nrCO
- 1.E-06 '5
2000
CU-
1.E-08
-
E 1600
1.E-1 0
1000-
500
0.000
0.003
0.006
0.009
1.E-12
0.012
Time (s)
Figure 4.26 - Temperature and OH concentration for constant-volume autoignition
simulations of iso-octane with 7% CO by energy, and of iso-octane with simulated
non-reactive CO (nrCO); Tinit = 875 K, Piit = 45 bar, X = 1.5.
119
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120
CHAPTER 5. ENGINE OPTIMIZATION
This chapter investigates some of the performance and efficiency optimization issues
relevant to modern internal combustion engines. Suggestions for maximizing low speed
torque and the best utilization of improved fuel octane quality for improved efficiency are
made and supported by processing the experimental data and applying a simplified
analysis.
5.1 PERFORMANCE ANALYSIS
This section investigates optimization of low speed torque by varying compression ratio,
and the effects of operating conditions on the fuel octane number requirement. The
results of the fuel octane number requirement analysis are used in the last section of this
chapter to analyze the benefits of increased fuel knock resistance.
5.1.1 Compression Ratio Optimization
A parameter that may be useful for the design of naturally aspirated engines is the
compression ratio for maximum low speed torque. The intake manifold pressure controls
the amount of air that is drawn into a spark ignition engine, and consequently the amount
of fuel energy that can be released. At low speeds, where intake manifold tuning is not
effective, the manifold pressure of naturally aspirated engines is limited to approximately
atmospheric pressure. Since the maximum amount of air-fuel mixture that can enter the
engine is limited, maximum torque is impacted by how efficiently the engine can turn the
air-fuel mixture into mechanical work. Compression ratio has two effects; increasing rc
increases maximum efficiency, but the spark retard required to avoid knock decreases
efficiency. The result is a tradeoff for which there is an optimum compression ratio.
Figure 5.1 illustrates this tradeoff for 1500 rpm. The top line is calculated from Eq. ( 3.2)
and represents the increase in maximum torque at MBT spark timing with increased
compression ratio. This torque benefit is only available with a very high ON fuel. The
combustion retard required to avoid knock, shown with the dashed line, was increased at
a rate of 3' CA per unit rc, which is consistent with the mean shown in Figure 3.53.
Combustion retard at the baseline 9.8:1 compression ratio was chosen to be 10" CA,
which matches that for TRF95. The bottom line is the decrease in torque from the
retarded combustion timing, calculated from Eq. ( 3.1). The middle solid line represents
121
the engine output NIMEP. It is the product of the increase from increased maximum
efficiency and the decrease from increased combustion retard:
NIMEP
(Eq.3.2).
NIMEPMBT
NIMEPIMEPMBT,9.8:1
NIMEP (Eq.3.1)
(5.1)
NIMEPMBT
-25
1.1
increased NIMEPMBT
from increased R-
1.05
20
CL increased combustion
retard to avoid knock
LLI
0
15
Ile,
Q
ME0.95
--
o
101%
4
product of increase from Rc
aed
0.9
E
and decrease from C.R.
NIMEP/NIMEPMBT
5
from increased C.R.
0.85
9
10
1
11
1
12
.
0
13
14
15
Rc
Figure 5.1 - Tradeoff between NIMEP and compression ratio. Values shown
approximate TRF95 fuel.
Under these conditions torque output peaks at a compression ratio of approximately
11.2:1. Table 5.1 shows how the compression ratio for peak NIMEP changes with fuel
type (which changes the combustion retard at the baseline 9.8:1 r,), and for values of 2, 3,
and 4' CA of combustion retard per unit rc, chosen to represent the spread indicated by
Figure 3.53. The trend with increased fuel octane for 3" CA per unit rc is similar to the
experimental values indicated by Figure 3.51. At typical fuel RON values of 90 and 95,
the compression ratio for peak NIMEP is relatively insensitive to the rate of combustion
122
retard increase with r. The average compression ratio for peak NIMEP at 1500 rpm is
approximately 11:1 for this engine. As engine speed increases, and the engine becomes
less knock limited, the compression ratio for maximum NIMEP would approach 14:1.
Table 5.1 - Calculated compression ratio for peak torque at 1500 rpm.
Fuel type
TRF85
TRF90
TRF95
TRF100
Corresponding
combustion retard
at 9.8:1 r, ('CA)
22
18
10
-5
Increase of combustion retard with
compression ratio ("CA per unit r,)
2
3
4
9.8:1
10.6:1
11.6:1
10.0:1
10.8:1
11.8:1
10.6:1
11.4:1
12.2:1
12.4:1
13.0:1
13.6:1
5.1.2 Fuel Octane Quality and Reformate Addition
For typical engine applications the shape of the low speed portion of the torque curve is
determined by the amount of combustion retard required to avoid knock. Maximum
combustion retard must be limited to avoid unsatisfactory performance, and to avoid
excessive exhaust temperatures and combustion instability.
Figure 5.2 shows contours of constant combustion retard on a chart of fuel RON vs.
reformate fraction at 1500 rpm, 11.6:1 rc, and 40% boost. They represent the decrease of
the octane number requirement of the primary fuel to maintain a given combustion retard
with increased reformate fraction. Previous work with PRFs indicates that reforming
30% of the fuel results in a decrease of about 20 ON in the PRF RON required to avoid
knock at MBT spark timing [14]. The PRF data in Figure 5.2 agrees with the previous
work and extends it to operating conditions with retarded spark. As discussed in Section
3.2.4 , hydrogen is not effective at slowing autoignition reactions for toluene. Thus the
decrease of combustion retard with reformate addition for high RON TRFs, which are
mostly toluene, is not as pronounced as it is for TRFs with moderate RONs, which have a
significant concentration of n-heptane. For TRF RONs between 90 and 100, adding 30%
reformate reduces ON requirement by about 10 ON.
The increase of fuel octane number requirement with boosted NIMEP for several values
of combustion retard at 1500 rpm are shown on Figure 5.3. As indicated by the
123
experimental data in Section 3.2.3 , PRFs are more sensitive to boost than TRFs. For a
real boosted application the inlet temperature would increase by 10-50 K, depending on
the effectiveness of the intercooler, so the increase of fuel RON requirement with boosted
NIMEP would be somewhat steeper. Experiments by Russ give the increase of octane
number requirement with inlet temperature to be 1 ON per 7 K [16].
I Iur
110
N umbers are
Numbers are
combustion retard
(CAD)
combustion retard
105
(~ C
I
05
z0
z
0
W100
00
U-
95
90-
95
--------
5
20
15
10
Reformate Fraction (%)
25
30
o
5
25
20
15
10
Reformate Fraction (%)
30
Figure 5.2 - Contours of constant combustion retard for varying reformate fraction
and fuel RON; X = 1.0, 1500 rpm, MAP for 40% boost. Data from Figure 3.45
(PRFs, left) and Figure 3.46 (TRFs, right).
Figure 5.4 shows the increase of fuel octane requirement with compression ratio at 1500
rpm, wide-open throttle. PRFs and TRFs behave similarly. The slope for most values of
combustion retard is 2-3 ON per unit compression ratio. This is somewhat less than that
predicted by Russ [16]. The differences may be partially due to the experiments in this
work being performed under retarded-spark conditions while the experiments in the
reference were performed only at MBT spark timing. They may also be partially due to
differences in engine configuration and inlet conditions.
124
10
U1
95-
100-
z0
z0
951-
90
Ix
a.
85
90[-
Numbers are
combustion retard
(CAD)
Numbers are
combustion re tard
(CAD)
9
11
12
10
Boosted NIMEP (bar)
13
14
9
08
"'V
10
11
14
13
12
Boosted NIMEP (bar)
Figure 5.3 - Contours of constant combustion retard for varying boosted NIMEP
and fuel RON; X = 1.0, 1500 rpm. Data from Figure 3.35 (PRFs, left) and Figure
3.36 (TRFs, right).
1uu
00
105
95
100
z
0z
w
0
L-
0.
90-
95
N
Numbers are
combustion retard
(CAD)
%A[I
10
10.5
11
11.5
Rc
12
12.5
13
85-
combustion retard
(CAD)
-0
10.5
11
11.5
Rc
12
12.5
13
Figure 5.4 - Contours of constant combustion retard for varying compression ratio
and fuel RON; X = 1.0, 1500 rpm, MAP = 1.0 bar.
125
5.2 EFFICIENCY ANALYSIS
As discussed further in the next section, improvement in fuel knock-resistance would
allow design parameters that are typically limited by knock to be extended. The simple
analysis presented in this section looks at the possible efficiency benefits of increasing
compression ratio, and of inlet boosting and engine downsizing.
5.2.1 Effect of Compression Ratio on Brake Efficiency
The effect of compression ratio on net efficiency is shown in the experimental results of
Section 3.1 . For vehicular applications, a more important parameter is brake efficiency,
which dictates fuel consumption. If the engine size is not changed, and engine friction
does not change significantly, then brake efficiency will stay proportional to the net
efficiency as compression ratio increases. The improvement in brake efficiency will then
be the same as the improvement in net efficiency. Also, since the engine size doesn't
change, the maximum torque output will increase as compression ratio increases. If,
however, the size of the engine is decreased so that maximum torque output remains
constant, a reduction in pumping work and friction causes brake efficiency to improve
more than net efficiency.
An illustration of this effect is shown in Table 5.2. For the analysis it was assumed that
mechanical friction normalized by engine size, FMEP, stays constant at 0.8 bar [28].
This assumption is reasonable because friction torque increases only slightly with load
compared to the total torque, and friction torque scales approximately with engine size.
Although maximum NIMEP would occur at a higher speed, the data for 1500 rpm
presented in Figure 3.48 was used because it was readily available and should scale
similarly to higher speeds. Maximum BMEP was estimated by subtracting the assumed
FMEP from the experimental maximum NIMEP:
BMEP = NIMEP - FMEP
(5.2)
As compression ratio increases BMEPax increases and the engine displacement is
decreased to keep maximum torque constant:
Torque(Nm) = 8 - BMEP(bar) -Displacement(L)
126
(5.3)
Displacementdownsized
BMEPmax,baseline
Displacementbaseline
BMEPmax,downsized
(54)
Since the engine displacement is reduced, mid-load BMEP must increase to keep midload torque constant:
BMEPmidoaddownsized
BMEPmid-load baseline
_
Displacementbaseline
Displacementdownsized
BMEPmax,downsized
BMEPmax,baseline
(55)
A mid-load BMEP of 2.6 bar at 1500 rpm was selected for the baseline 9.8:1
compression ratio because it is representative of the mid-load cruising condition and is
influential on the overall vehicle efficiency [27]. Mid-load NIMEP is calculated by
adding FMEP to the downsized BMEP. Net efficiency is interpolated from Figure 3.6.
Brake efficiency is then estimated by multiplying the net efficiency by the ratio of BMEP
to NIMEP.
(5.6)
=
ne - Bbrake
77brak
NIME
Figure 5.5 compares the improvement in brake efficiency with downsizing, calculated as
described above, to the improvement in brake efficiency without downsizing, calculated
by assuming constant mechanical efficiency. Downsizing the engine to maintain a
constant maximum torque output with increased rc increases the efficiency benefit by
about 60% compared to the case without downsizing.
Table 5.2 - Effect of downsizing with increased r, on brake efficiency.
R
9.83
11.59
13.40
Max
Max
NIMEP
Max
Max
BMEP
Midload
BMEP
Midload
NIMEP
(bar)
(bar)
(bar)
(bar)
10.16
10.74
11.01
9.36
9.94
10.21
2.60
2.76
2.84
3.40
3.56
3.64
Net
efficiency
Brake
efficiency
Increase
in brake
efficiency
(%)
0.289
0.302
0.309
0.221
0.234
0.241
0
6.1
9.3
127
12
,a- Without Downsizing
I
-~10
VVILII
LJUW"ZI1,LiII
8
W
1
_
4-
02
9
10
12
11
13
14
Rc
Figure 5.5 - Estimated increase of mid-load brake efficiency with r,, with and
without downsizing the engine to maintain constant maximum torque output; 1500
rpm, X = 1.0, 2.6 bar baseline BMEP.
5.2.2 Effect of Boost on Brake Efficiency
Turbocharging an engine increases its maximum BMEP. If engine size is held constant,
this can be used to increase the torque output of the engine without significantly affecting
efficiency (if compression ratio is not changed). However, if the engine is downsized to
keep torque output constant, as described for increased compression ratio in the previous
section, decreased friction, pumping work, and heat transfer lead to an increase in brake
efficiency.
The effect of boosting and downsizing is illustrated in Table 5.3. As in the previous case,
FMEP was assumed to be constant at 0.8 bar, baseline mid-load BMEP is assumed to be
2.6 bar, and mid-load BMEP is scaled with maximum BMEP. A baseline maximum
BMEP of 10.4 bar was chosen to match the performance of modem engines [4]. The
increase in net efficiency is interpolated from Figure 3.6. As in the previous section,
brake efficiency is estimated by multiplying the net efficiency by the ratio of BMEP to
NIMEP. The improvements in net and brake efficiencies are plotted in Figure 5.6. The
128
improvement in net efficiency is due to decreased pumping work and heat transfer, the
additional improvement in brake efficiency is due to reduced friction.
Table 5.3 - Effect of boosting and downsizing on mid-load brake efficiency.
NIMEP
boost
bevsl
level (%)
0
20
40
Midload
BMEP
Max
BMEP
baE
Max
NIMEP
NIaEP
(bar)
(bar)
(bar)
Net
eff.
Brake
eff.
(bar)
Increase
in brake
efficiency
(%)
0.289
0.302
0.312
3.40
3.96
4.52
2.60
3.16
3.72
10.40
12.64
14.88
11.20
13.44
15.68
Midload
NIMEP
0.221
0.241
0.256
0
9.0
16.2
20
Efficiency
-Net
---
Brake Efficiency
16
C
12
0
8
C)
4
0
0
10
20
30
40
50
NIMEP Boost Level (%)
Figure 5.6 - Estimated increase of mid-load net and brake efficiencies with boosting
and downsizing; 1500 rpm, X = 1.0, 2.6 bar baseline BMEP.
129
5.3 IMPLICATIONS FOR ENGINE DESIGN
Modem engines are designed to maximize performance and fuel efficiency within the
constraints of available fuel quality. The experimental results presented in this work
indicate that an improvement of 10 ON or more can be realized by converting a portion
of the fuel entering an engine to a hydrogen rich reformate. Consequently, it is an
interesting exercise to analyze what the efficiency benefits would be for a 10 ON
improvement in fuel knock resistance, whether it be through on-board reforming, or
through improved fuel refining.
Assuming that gasoline performs in between PRFs and TRFs, Figure 5.3 implies that
boosting NIMEP by 40% while maintaining constant combustion retard would require an
improvement of fuel quality of about 6 ON. If inlet temperature increases by about 30 K,
it would bring the increased octane requirement to about 10 ON [16]. Alternatively,
Figure 5.4 implies that increasing fuel quality by 10 ON would allow compression ratio
to be increased from 9.8:1 to 13.4:1 without increasing the amount of combustion retard
required to avoid knock.
The estimations made in this chapter indicate that with the improved fuel quality,
increasing compression ratio could potentially improve brake fuel conversion efficiency
by about 9%. Adding a turbocharger, which would have a somewhat higher cost, could
improve brake fuel conversion efficiency by up to 16%. To maximize efficiency benefit
given an improvement in fuel octane quality over today's levels, this analysis implies that
inlet boosting and engine downsizing is a better strategy than increasing compression
ratio. Since the calculations presented required many assumptions and simplifications, a
more accurate engine performance analysis may be necessary to confirm the magnitude
of the efficiency benefits.
130
CHAPTER 6. CONCLUSIONS
*
Mid-load net efficiency improvement from increased air-fuel ratio peaks at about
10% relative to stoichiometric values. About 2/3 of the improvement is from the
increased ratio of specific heats for lean mixtures and reduced heat transfer; the rest is
from reduced pumping losses. The air-fuel ratio for peak efficiency increases with
increasing compression ratio. Relative net efficiency improvement from increasing
load is about 7% per bar NIMEP at mid-load. About 75% of the improvement is
from reduced pumping losses and 25% is from heat loss becoming a smaller portion
of the overall charge energy. At a compression ratio of 9.8:1, relative net efficiency
improvement is about 2.5% per unit compression ratio. Efficiency peaks at about
14:1 with a maximum benefit of 6-7%. Efficiency improves more with compression
ratio at high speeds and loads due to the reduced importance of heat loss.
* A new combustion phasing parameter, termed "combustion retard", has been
developed. It is the location of 50% mass fraction burned minus the location of 50%
mass fraction burned for MBT spark timing. It represents the crank angle that the
center of the combustion event has been shifted from the timing for maximum torque.
Loss of indicated torque correlates well to a diagonally asymptotic relationship with
combustion retard for the wide range of operating conditions considered.
* The knock behavior of three fuel types was investigated. Primary reference fuels are
mixtures of iso-octane and n-heptane or tetraethyl lead that are used in the ASTM
octane rating procedures. Toluene reference fuels are mixtures of toluene and nheptane that are used to better represent the aromatic content and sensitivity to
operating conditions of gasoline. The octane number rating of both of these fuel
types can be adjusted by adjusting the mixture composition. Two formulations of
unleaded test gasolines were used to represent regular automotive fuel. The knock
behavior of PRFs and TRFs is similar for changes in compression ratio, but PRFs are
more sensitive to boosted inlet pressure, air-fuel ratio, and fuel reformate than TRFs.
Test gasolines behave about halfway between PRFs and TRFs.
* The combustion retard required to just avoid knock for PRFs increases with air-fuel
ratio, with the increase becoming more severe at higher compression ratios. For
TRFs, required combustion retard decreases slightly with air-fuel ratio for all
131
compression ratios. PRFs, which require about 5' CA of combustion retard per bar
NIMEP, need about three times as much combustion retard as TRFs when boosted to
achieve the same NIMEP. Both fuel types require an average of about 30 CA of
combustion retard per unit of increased compression ratio. The standard deviation of
the data is about 50% of the total effects. The wide spread is mainly due to the nonlinear nature of the knock phenomenon, but is also affected by the accuracy of
audible knock detection and changing atmospheric conditions.
" The combustion retard required to just avoid knock decreases by about 20 CA per 3%
reformed fraction for PRFs. For TRFs with low alkane content reformate addition is
less effective. Constant volume iso-octane autoignition modeling shows that
hydrogen converts hydroxy radicals to hydrogen radicals, which are not as effective
at initiating the chain branching reaction sequence. In the toluene oxidation
mechanism investigated, most of the initiation and propagation reactions involving
the hydroxy radical have an analogous reaction with the hydrogen radical with similar
rate constants, so the overall reaction rate is not affected.
*
Detailed chemical kinetics mechanisms were combined with a cylinder pressure
based end-gas modeling methodology. It was found that high accuracy in the
determination of initial conditions is required for consistent results. Calibration of
baseline initial temperature is required to match experimental data. The model
successfully predicted the response of PRFs to compression ratio and air-fuel ratio
and the response of TRFs to boost. The difference between the response of PRFs
and TRFs to air-fuel ratio was also captured. Simulations with reaction kinetics
disabled show that due to a higher ratio of specific heats, TRFs are more sensitive to
the pressure ratio through which the end-gas is compressed than PRFs. This effect
helps to explain the different responses of PRFs and TRFs to boosted lean operation.
*
Reforming 30% of the fuel entering an engine decreases the required fuel octane
number by 10 ON or more depending on fuel composition. This improvement would
allow the compression ratio of an engine to be raised from 9.8:1 to approximately
13.4:1 without increasing combustion retard, or it would allow the engine to be
boosted to a 40% higher NIMEP. A simplified analysis indicates that increasing
compression ratio and downsizing the engine to maintain constant maximum torque
would increase mid-load fuel efficiency by about 9%. Boosting and downsizing
would increase mid-load fuel efficiency by about 16%.
132
REFERENCES
[1]
Heywood, J.B., Internal Combustion Engine Fundamentals,McGraw Hill Inc.,
New York, 1988
[2]
Pan, J., et al., "End-Gas Inhomogeneity, Autoignition and Knock," SAE 982616
[3]
Tanaka et al., "Two-stage ignition in HCCI combustion and HCCI control by
fuels and additives," Combustion and Flame, Volume 143, pp219-239, 2003.
[4]
Chon, D.M., and Heywood, J.B., "Performance Scaling of Spark Ignition
Engines: Correlation and Historical Analysis of Production Engine Data," SAE
2000-01-0565
[5]
Taylor, C., The Internal Combustion Engine in Theory and Practice, M.I.T. Press,
Cambridge, MA 1985
[6]
Goldwitz, J.A., and Heywood, J.B., "Combustion Optimization in a HydrogenEnhanced Lean-Bum SI Engine," SAE 2005-01-0251
[7]
Hisato Hirooka, Sachio Mori and Rio Shimizu, "Effects of High Turbulence Flow
on Knock Characteristics," SAE 2004-01-0977
[8]
Topinka, J., Knock Behavior of a Lean-Burn, H2 and CO Enhanced, SI Gasoline
Engine Concept ,M.S. Thesis, MIT, May 2002
[9]
Glassman, I., Combustion, Academic Press Inc., California, 1996
[10]
Ivanic, Z. et al., "Effects of Hydrogen Enhancement on Efficiency and NOx
Emissions of Lean and EGR Diluted Mixtures in a SI Engine," SAE 2005-010253
[11]
Natkin, R., Ford Motor Company, Private Communication, June 2003.
[12]
Green, R., and Pearce, S., "Alternative Transport Fuel," Energy World Journal,
pp. 8-11. October 1994.
[13]
Tang, X., Heffel, J., et al., "Ford P2000 Hydrogen Engine Dynamometer
Development", SAE 2002-01-0242
133
[14]
Topinka, J.A. et al., "Knock Behavior of a Lean-Burn, H2 and CO Enhanced, SI
Gasoline Engine Concept," SAE 2004-01-0975
[15]
Gruden, D., and Hahn, R., "Performance, Exhaust Emissions and Fuel
Consumption of a IC Engine with Lean Mixtures," I Mech E Publication
(C111/79), 1979.
[16]
Russ, S., "A Review of the Effect of Engine Operating Conditions on Borderline
Knock," SAE 960497
[17]
Muranaka, S., Takagi, Y., and Ishida, T., "Factors Limiting the Improvement in
Thermal Efficiency of S.I. Engine at Higher Compression Ratio," SAE 870548
[18]
Kalghatgi, G., Shell Global Solutions, Private Communication, April 2004.
[19]
Burluka, A., et al., "The Influence of Simulated Residual and NO Concentrations
on Knock Onset for PRFs and Gasolines," SAE 2004-01-2998
[20]
Kalghatgi, G., "Auto-ignition quality of practical fuels and implications for fuel
requirements of future SI and HCCI engines," SAE 2005-01-0239
[21]
Emdee, J., Brezinsky, K., and Glassman, I., "A Kinetic Model for the Oxidation
of Toluene Near 1200K," J. Phys. Chem. 1992, 96, 2151-2161
[22]
Pitz, W., et al., "Chemical Kinetic Study of Toluene Oxidation Under Premixed
and Nonpremixed Conditions," LLNL report UCRL-CONF-201575, 2003
[23]
Curran, H. J., et al., "A Comprehensive Modeling Study of iso-Octane
Oxidation," Combustion and Flame, Volume 129, pp253-280, 2002.
[24]
Curran, H. J., et al., "A Comprehensive Modeling Study of nHeptane Oxidation,"
Combustion and Flame, Volume 114, pp149-177, 1998.
[25]
Tomohiro Shinigawa et al., "Effects of Hydrogen Addition to SI Engine on Knock
Behavior," SAE 2004-01-1851
[26]
Scire, J. J., et al., "Flow Reactor Studies of Methyl Radical Oxidation Reactions
in Methane-Perturbed Moist Carbon Monoxide Oxidation at High Pressure With
Model Sensitivity Analysis," International Journal of Chemical Kinetics, Volume
33, pp75-100, 2001
134
[27]
Leone, T., Ford Motor Company, Private Communication, April 2005.
[28]
Sandoval, D., and Heywood, J.B., "An Improved Friction Model for SparkIgnition Engines," SAE 2003-01-0725
135
(this page intentionally left blank)
136
APPENDIX A MODIFIED PISTON DIMENSIONS
137
ref 6.64
71.23
ref 1.36
36
0 33100
137
ref 16,08
56.12
------
-
-
--5
-
- --
--
--
--
-
35.9
30
87~
-- ----
12:1 CR Piston
Mike Gerty
Oct 22, 20 3
138
ref 6.64
71,23
ref 1.36
36
0 33.90
13.7
ref 16 08
56 12
--
077.51
------
----
-----
35.9
5,4
87
14:1 RC Piston
Mike Gerty
Oct 22, 2003
139
(this page intentionally left blank)
140
APPENDIX B ADDITIONAL CHARTS FOR EFFECTS OF
AIR-FUEL RATIO ON KNOCK
141
35
C-
Torque Loss
30
1 20%
25
-
-g-PRF100
-*- PRF95
UPRF90
-6-PRF85
20 d 10%
-+-PRF80
C
15
0
0
.0
-
10 - 3%
5
1%
0
-5
1.1
0.9
1.3
1.5
1.7
Lambda
Figure B. 1 - Effect of X on combustion retard to just avoid knock for PRFs; r. =
9.8:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
.
35
a
I-
Torque Loss
-*- TRF100
30- 20%
-4-TRF95
25-
-U-TRF90
0
m
U
C
so
0
--
20- 10%
C
so
-+-TRF80
0
I-
TRF85
15-
(U
4-.
(U
10 - 3%
0
(I)
5
1%
.0
E
0
C-)
0A
-5
0.9
'PO
1.1
1.3
1.5
1.7
Lambda
Figure B. 2 - Effect of X on combustion retard to just avoid knock for TRFs; r, =
9.8:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
142
35
Torque Loss
0
30 - 20%
C.)
25 -
-#- PRFIIO
-X-PRFIO5
*PRF100
20 - 10%
-+-
PRF95
-
PRF90
15-
10- 3%
0
X.
5 1%
E
0
- --
0
- --
- --
--
- ---
--
-- -
-5 1.1
0.9
1.7
1.5
1.3
Lambda
Figure B. 3 - Effect of A on combustion retard to just avoid knock for PRFs; re =
11.6:1, 1500 rpm, MAP at X> 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
35
Torque Loss
-*- TRF100
30 - 20%
-+-TRF95
25 -
-6-TRF90
-4-TRF85
20- 10%
15 0
U
10- 3%
5
1%
0 ------------------------5
-t
0.9
1.1
1.3
1.5
1.7
Lambda
Figure B. 4 - Effect of X on combustion retard to just avoid knock for TRFs; re =
11.6:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
143
35 .oqueLs
Torq ue Loss
I
30
--- PRF115
20%
-*- PRF1 10
25 -
-X- PRF105
-PRF00
20 1 10%
-+PRF95
15
-
10 - 3%
- - - - - - - -
-- - - - - - - -
0
1%
5
0 --
'
-5
1.1
0.9
1.3
1.5
1.7
Lambda
Figure B. 5 - Effect of X on combustion retard to just avoid knock for PRFs; r, =
13.4:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
35
Torque Los
-0--TRF90
30 - 20%
0
-4-TRF95
25 -
-i-
20 - 10%
-X-TRF105
TRF100
15 10-
3%
5
1%
0
0
--
---
-5
0.9
1.1
1.3
Lambda
1.5
1.7
Figure B. 6 - Effect of X on combustion retard to just avoid knock for TRFs; r, =
13.4:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at stoichiometric
unboosted WOT.
144
35
T
Torque Loss
-
30 - 20%
-
+- PRF95
-TRF95
-- +-UTG96
25
E
C
20 - 10%
--..
,
C
15
0
10 - 3%
.
0
E
.0
0
5 1%
0
-5
0.9
1.1
1.3
1.5
1.7
Lambda
Figure B. 7 - Effect of X on combustion retard to just avoid knock for UTG96 fuel;
r,= 13.4:1, 1500 rpm, MAP at X > 1.0 boosted to match MBT NIMEP at
stoichiometric unboosted WOT.
145
(this page intentionally left blank)
146
APPENDIX C ADDITIONAL CHARTS FOR EFFECTS OF
BOOST ON KNOCK
147
35
I-
Torque Loss
-)*-PRF105
30 - 20%
-4-PRF100
0
25 -
20 d 10%
0
--
zl"MOoe
PRF95
-0-PRF90
101
3%
0
5E
0
1%
----
0-
---
-----------
-5
8
9
10
11
12
13
14
15
NIMEP (bar)
Figure C. 1 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for PRFs; r,= 9.8:1, X = 1.0, 1500 rpm.
,
35
I-
Torque Loss
-*-TRF100
30- 20%
-+-TRF95
25 -
-U-TRF90
-0-TRF85
20- 10%
0
15-
10 - 3%
0
1%
5-
E
0
0-
----
--
- ---
- ---
---
--
-5
8
9
10
11
12
13
14
15
NIMEP (bar)
Figure C. 2 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for TRFs; re = 9.8:1, X = 1.0, 1500 rpm.
148
35
I Torque Loss
--
PRF100
30 - 20%
--
PRFI5
25 -
-N-
PRF110
-PRF115
20 - 10%
/
0
15 E
~0
0
10
-
--
PRF120
3%
5- 1%
0-
-5 i
9
1
1
10
11
1
1
16
15
14
13
12
NIMEP (bar)
Figure C. 3 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for PRFs; r,= 11.6:1, X = 1.0, 1500 rpm.
35
I-
-
Torque Loss
-*-Toluene
-+-
30- 20%
TRF105
-*-TRF100
25-
-4-TRF95
0
20 - 10%
0
15-
E
0
10 - 3%
0
1%
5
0
-5
9
10
11
12
13
14
15
16
NIMEP (bar)
Figure C. 4 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for TRFs; r,= 11.6:1, X = 1.0, 1500 rpm.
149
...............
35.
..
in25
Rc=13.4:1
Rc= 11.61
-... Rc=9.8:1
-.
--....
0
'a
C-
20.....
1 5 ....
- ....
..
..
........
1 0 ..
E
0
5.. .
0-
90
-....
100-
10
110
RON
9
1
0
12
13
14
Boosted NIMEP (bar)
Figure C. 5 - Effects of boosted NIMEP and PRF fuel RON on combustion retard to
just avoid knock; X = 1.3, 1500 rpm.
.............-.....
35 ................
3 0 .............
.....
........
2..
-.
Rc= 1 3 .4 :1
Rc=11.6:1
Rc=9.8:1
0
0)
-o
20 .
.....
--...
0
L)
E
..
-5 ....-..
-..
.....
0 ...
105
12014
120 9
10Boosted NIMEP (bar)
Figure C. 6 - Effects of boosted NIMEP and TRF fuel RON on combustion retard
to just avoid knock; X = 1.3, 1500 rpm.
150
35
I-
I Torque Loss
-X-PRF105
-h-PRF100
30 -120%
+- PRF95
25 -
//
-U-
20 -i 10%
PRF90
-S- PRF85
a
15 -
//
10 1I3%
U-
5
E
0
1%
0 -------------------5
9
8
11
10
NIMEP (bar)
12
13
Figure C. 7 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for PRFs; r,= 9.8:1, k = 1.3, 1500 rpm.
35
I-
30-
Torque Loss
-e- TRF100
20%
-+-TRF95
-U-TRF90
0
*E 25-
-+-TRF85
C
2015 -
10- 3%
.0
I-
0
5
1%
0
E
----
0 -5
-
-----
9
10
11
12
13
14
NIMEP (bar)
Figure C. 8 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for TRFs; r,= 9.8:1, X = 1.3, 1500 rpm.
151
. Torque Loss
35
C-
-
PRFI15
30 - 20%
-*- PRF110
25 -
-<-PRF105
-*-PRF100
20-
1U0/
-4-PRF95
0
15 -
E
10- 3%
1%
5
0
.0
0 --------
---------------
-5
9
10
11
12
13
14
NIMEP (bar)
Figure C. 9 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for PRFs; re = 11.6:1, X = 1.3, 1500 rpm.
35
30
-
25
-
Torque Loss
-M-TRF90
20%
-+-TRF95
-*-TRF100
0
-)-TRF105
20 - 10%
0
E
0
15
-
10 - 3%
5
1%
A-
U
- - -
- - - --
0 -- - - - - - -5
9
10
11
12
13
14
NIMEP (bar)
Figure C. 10 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for TRFs; r,= 11.6:1, X = 1.3, 1500 rpm.
152
.
It
Torque Loss
-PRFI15
30 - 20%
I0
25
-*-PRFIIO
7/M
-
-N-PRF105
C:
V
-*-PRF100
20 - 10%
7
15
-
10
- 3%
5
1%
0
0
-5
'
11
10
9
14
13
12
NIMEP (bar)
Figure C. 11 - Increase of combustion retard to just avoid kn )ck with increased
NIMEP from boosting for PRFs; r,= 13.4:1, X = 1.3, 1500 rpi 1.
35
L-
0
Torque Loss
--
TRF95
30- 20%
-6-TRF100
25-
-)-TRF105
20- 10%
I-
0
10d 3%
0
(U
5
1%
An -I-J------------------------5
9
10
11
12
13
14
NIMEP (bar)
Figure C. 12 - Increase of combustion retard to just avoid knock with increased
NIMEP from boosting for TRFs; re = 13.4:1, X = 1.3, 1500 rpm.
153
(this page intentionally left blank)
154
APPENDIX D ADDITIONAL CHARTS FOR EFFECT OF
FUEL REFORMATE ON KNOCK
155
35
16
-+-PRF100
C.R.
30 15
C-
25
-U--PRF90
-
C.R.
14
w
201
0
PRF95
C.R.
-
-
4- PRFIOO
-
--
-
a- PRF90
NIMEP
-,
130.i
1 IL
.D 10
12
-
0
-- , '
. .
- - - --
- -.
,
PRF95
NIMEP
NIMEP
-
5-
E
0
0-5
10
0
5
10
15
20
30
25
35
Reformate Fraction (%)
Figure D. 1 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for PRFs; r, = 9.8:1, X = 1.0,
1500 rpm, MAP for 40% boost (NIMEPMBT 1.4 times unboosted NIMEPMBT).
35
30
C0
w
16
-
TRF100
---
C.R.
-
-- 15
-+-TRF95
C.R.
25
-U-TRF90
-- 14
20
C.R.
--
C
0
TRF85
C.R.
15
-
10
-
----------
--
- -- -
- -12
-
4- TRF100
NIMEP
z
-
TRF95
--
NIMEP
50
E
0
- - 11
0-
110
-5
-
0
5
10
15
20
25
30
10
-
-
TRF90
NIMEP
- .0-
TRF85
NIMEP
35
Reformate Fraction (%)
Figure D. 2 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for TRFs; r, = 9.8:1, X = 1.0,
1500 rpm, MAP for 40% boost (NIMEPMBT 1.4 times unboosted NIMEPMBT).
156
. 16
35
I-
30 -
-~--PRF110
C.R
-+4-PRF105
- 15
0
-
-
S..-
-
10 25 -*
0
C
C
0
-
10
-
PRF100
-
20U-
C.R.
PRF95
C.R
PRF115
+
- 13
z
- --
NIMEP
- 12
0
5
E
0
C.R
-14
15
PRF115
C.R
---
-- - - - - - - - - -
0
r::
-
*- -PRFIIO
-
-x-
NIMEP
PRFIO5
NIMEP
-PRFIOO
- --
-5 i
0
I5
10
15
25
20
30
NIMEP
1IV
PRF95
35
NIMEP
Reformate Fraction (%)
Figure D. 3 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for PRFs; r, = 13.4:1, X =
1.0, 1500 rpm, MAP for 40% boost (NIMEPMBT 1.4 times unboosted NIMEPMBT). 16
35
----
Toluene
C.R.
30
.. - - - - - . x-
- - - -
-- +-TRF105
15
C.R.
250-
-
-+--TRF100
C.R.
14
20
U
---
13 Iw
15
E
0
TRF95
C.R.
-
--
z
10
12
0
oluene
NIMEP
-
X- TRF105
NIMEP
5
11
-
TRFIOO
-
NIMEP
0
---
'
-5
0
5
10
15
20
25
30
10
TRF95
NIMEP
35
Reformate Fraction (%)
Figure D. 4 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for TRFs; r, = 13.4:1, X =
1.0, 1500 rpm, MAP for 40% boost (NIMEPMBT 1.4 times unboosted NIMEPMBT)157
- ..
...................
-- ..
4
Rc=13.4:1
R c= 1 1 .6 :1
Rc=9.8:1
0 ...--....
R30
eo
3 3525-
.....
20
15
0
-..-....-...
10
20
Reformate Fraction (%)
.....
30 ..... .....100
90
RON
Figure D. 5 - Effects of reformed fuel fraction and PRF fuel RON on combustion
retard to just avoid knock; X=1.3, 1500 rpm, MAP set to match NIMEPMBT at
stoichiometric unboosted
WOT.
0
8
Rc=13.4:1
......Rc=1 1.6:1
- Rc=9.8:1
( 35 -.........
30 .-...
25 0
.....
101E
0~2
R.N
......
Reformate Fr....n.(%)..
Figue
ffecsD.6
ofrefomed
uelfracion nd TF
retard t........
~0
s... i k..k
,.....,.A.s.to
.. nb..sted .............
i
st h5mt
5
8....8
....
ful.R.......b.ti.
m tc.....IM
*v
.=. P BT a
35
30
-
-- *--PRF95
C.R.
-- U-PRF90
11
C.R.
25:
-*-PRF85
20 1
C.R.
---- PRF80
A
..
10
0
15
-
10
-
w
a.
~-
a
- - - - - -- --
-
-
-o-
C.R.
PRF75
C.R.
PRF95
NIMEP
-
9
5
E
0
---
-PRF90
-
NIMEP
-
--
-
-PRF80
0 -5
NIMEP
I
0
PRF85
I
5
i
I
'
I
25
20
15
10
Reformate Fraction (%)
30
8
NIMEP
35
- -x-
PRF75
NIMEP
Figure D. 7 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for PRFs; r, = 9.8:1, X = 1.3,
1500 rpm, MAP set to match NIMEPMBT at stoichiometric unboosted WOT.
35
-- +-TRF95
C.R.
30
-
25
-
0
20
-
a
15
-
I-
-
11
-U- TRF90
-e-
IL
U~ 10
w
- - - --
- -----
- -
--
-+-TRF80
C.R.
TRF75
--C.R.
- -o- TRF95
- -- - -
NIMEP
-9
- --
5-
E
0
C.R.
TRF85
C.R.
TRF90
NIMEP
-
<- TRF85
0-
NIMEP
- +-'- TRF80
I-8
-5
0
5
10
15
20
25
Reformate Fraction (%)
30
35
NIMEP
-
--
-TRF75
NIMEP
Figure D. 8 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for TRFs; r, = 9.8:1, X = 1.3,
1500 rpm, MAP set to match NIMEPMBT at stoichiometric unboosted WOT.
159
35
PRF100
C.R.
PRF95
C.R.
PRF90
C.R.
-+I-
30 -
11
--
0 .- 0
25 1
0
-
20 -
-10
0~ 1
is -
-0
0
-+-PRF100
NIMEP
-PRF95
-
NIMEP
--p
0
9
'0'
5 11
0
C.R.
PRF80
C.R.
-
10-
(0
E
PRF85
E
--
-- ---------- --
-PRF90
-
NIMEP
- +
8
-5
0
5
10
15
20
25
NIMEP
35
30
PRF85
-
PRF80
--
NIMEP
Reformate Fraction (%)
Figure D. 9 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for PRFs; r, = 11.6:1, X =
1.3, 1500 rpm, MAP set to match NIMEPMBT at stoichiometric unboosted WOT.
35
-+-TRF100
C.R.
30 ~
25
-
. -. . .
- - -.-. .
.
11
-- 'TRF95
C.R.
-4-TRF90
3
0
0
15,a
10
w
-+--TRF80
C.R.
z
-
-
-
-9
5
-
+~
TRF95
-
a-
-TRF90
NIMEP
0 -
NIMEP
- .0-
-5
TRF100
NIMEP
-
--- - - -- - -
E
0
C.R.
TRF85
C.R.
---
20
,8
10
5
10
15
20
25
Reformate Fraction (%)
30
35
TRF85
NIMEP
- +*- TRF80
NIMEP
Figure D. 10 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for TRFs; r, = 11.6:1, X =
1.3, 1500 rpm, MAP set to match NIMEPMBT at stoichiometric unboosted WOT.
160
-
35
-- +- PRF105
C.R.
I-
30 -
0
25 -
11
U
--
PRFI0
C.R.
- PRF95
00
C.R.
0
0,~
20 -
PRF90
C.R.
S
OPRF85
-
10
a.
w
C.R.
-
10 -
-x- 'PRF105
NIMEP
0
9
5-
PRF100
- -A-
NIMEP
-PRF95
- --
E
0
- - - - - - - - - - - -
0 -- - - - - - - - -
NIMEP
-
-PRF90
NIMEP
8
-5
0
10
5
35
30
25
20
15
-PRF85
-
NIMEP
Reformate Fraction (%)
Figure D. 11 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for PRFs; r, = 13.4:1, X =
1.3, 1500 rpm, MAP set to match NIMEPMBT at stoichiometric unboosted WOT.
35
I-
30
--. .-
C
-- ) -TRF105
11
. - . --.
+ - --
--
25
-
...
.
...
. -. -. - - -
A
-
0
20
10
E
15
a.
w
-
0
.- - - - - -
-
-- -
.
,-
z
10
C.R.
-4-TRF95
C.R.
-U-TRF90
C.R.
-+-TRF85
C.R.
-
-x- TRF105
NIMEP
0
@1
TRF100
C.R.
9
-
-
5-
TRFIOO
NIMEP
-
--
-
a-
TRF95
NIMEP
0-5
0
5
10
15
20
25
Reformate Fraction (%)
30
35
TRF90
NIMEP
8
-0
. TRF85
NIMEP
Figure D. 12 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for TRFs; r, = 13.4:1, X =
1.3, 1500 rpm, MAP set to match NIMEPMBT at stoichiometric unboosted WOT.
161
35
+- UTG96
0 30 -1C.R.
- -- UTG91
- 2C.R.
25-
-
,-
-10
-.-
UTG96
NIMEP
UTG91
o
-NIMEP
15
~-
~X20
-
2W
,
w10 -
-- 9
5
.0
E
0
0 ---
- ---
--
- --
5
10
15
--
--
--
--
--
-
-5
8
0
20
25
30
35
Reformate Fraction (%)
Figure D. 13 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for UTGs; re = 11.6:1, x =
1.3, 1500 rpm, MAP set to match NIMEPMBT at stoichiometric unboosted WOT.
35
Q
---
UTG96
C.R.
-1 -
UTG96
30 -
o -NIMEP
25
20
-
-- 10
00
15 -
00
U)
5z
0--9
05-
E
0
---------------------
0----------
-5
-8
0
5
10
15
20
25
30
35
Reformate Fraction (%)
Figure D. 14 - Decrease of combustion retard to just avoid knock and associated
increase in NIMEP for increased reformed fuel fraction for UTG96; r, = 13.4:1, x =
1.3, 1500 rpm, MAP set to match NIMEPMBT at stoichiometric unboosted WOT.
162
APPENDIX E ADDITIONAL CHARTS FOR EFFECT OF
COMPRESSION RATIO ON KNOCK
163
-
35
I 30
11
---- PRF105
C.R.
-
-- *-PRF100
C.R.
- 10
5.
25-
-2015
-9
10 -
0
E
0
i
-- e--PRF90
C.R.
0.
-
- --
-
--
-
- - -- - --
-
-x- PRF105
-
--
NIMEP
-8 2
-
PRF95
C.R.
-PRFIOO
NIMEP
5-7
-
--
NIMEP
0-
-PRF90
NIMEP
-6
-5
10
9
11
12
13
PRF95
14
Rc
Figure E. 1 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for PRFs; X = 1.3, 1500 rpm, MAP set to match
NIMEPMBT at stoichiometric unboosted WOT.
-
35
11
C.R.
30 10
a
25
-
20
-
TRF90
C.R.
-,+-TRF85
C.R.
1510
-+-TRF95
C.R.
---
I0
M
-- *-TRF100
------
-
-- - -
-------
-
-
4- TRF100
NIMEP
8 E
-
-
0
4- TRF95
NIMEP
50
7
-
a- TRF90
NIMEP
-
.-
06
-5
9
10
11
12
13
TRF85
NIMEP
14
Rc
Figure E. 2 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for TRFs; X = 1.3, 1500 rpm, MAP set to match
NIMEPMBT at stoichiometric unboosted WOT.
164
'
.
35
)K.
30 -
-*-PRF110
C.R.
--
-
,
k
PRF105
C.R.
. . .. . . . .
25 -
0
11
-
PRFI
C.R.
-
20
0
-
-- *-PRF95
C.R.
0
E-
15
-
(0
E
0
PRF110
- -X-
NIMEP
-
- -PRF105
5-
NIMEP
7
E
-- -PRF100
-
0-
C.)
z
8
10
00
-5
- - - - - - - - - - - - - - -
-- --
-
I
9
10
I
I
-
I
- - -PRF95
6
I
NIMEP
14
13
12
11
NIMEP
Rc
Figure E. 3 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for PRFs; X = 1.6, 1500 rpm, MAP set to match
NIMEPMBT at stoichiometric unboosted WOT.
' 11
35
-+-TRF100
C.R.
30
-
C
25
-
-
-.
. -
-
-.
-
10
-
-'&--TRF90
C.R.
M
at
20
-- +-TRF95
C.R.
94
-
0W
8 00
15-
-+-TRF85
C.R.
-
4- TRF100
NIMEP
0
10
-
-
-
TRF95
NIMEP
0
E
0
5-
7
0-5 4
9
6
10
12
11
13
-
--
-
.o- TRF85
TRF90
NIMEP
NIMEP
14
Rc
Figure E. 4 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for TRFs; X = 1.6, 1500 rpm, MAP set to match
NIMEPMBT at stoichiometric unboosted WOT.
165
-
35
30
14
PRF115
C.R.
PRF110
C.R.
-
-
---13
25
-
CP
20
-PRFI
--
I
-
12
~
--
--
0
C.R.
CL
15
-
-
I-
10
4'
11
-
PRF95
C.R.
PRF115
NIMEP
E
0
05
C.R.
PRFIO9
-
5-
--
PRF110
NIMEP
10
0-
- -X-
PRF105
- --
-PRF100
NIMEP
'
-5
9
9
10
11
12
13
NIMEP
14
- -o-
PRF95
NIMEP
Rc
Figure E. 5 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased re for PRFs; X = 1.0, 1500 rpm, MAP for 20% boost
(NIMEPMBT
1.2 times unboosted NIMEPMBT).
35
30
-
14
--
TRF105
C.R.
-
25
-
20
-
13
I.----.
---
TRF100
C.R.
---
TRF95
C.R.
. . . . . . ..
-
12'g
0
--U-TRF90
.0
C.R.
15-
-
-TRF105
A-
-TRF100
-
E
0
NIMEP
-112
10-
-
0
NIMEP
5 -
0
10
4 -- - - - - - - - - - - -- - - - - - - - - -9
-5
9
10
11
12
13
- --
-TRF95
NIMEP
-
-- -TRF90
NIMEP
14
Rc
Figure E. 6 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for TRFs; X = 1.0, 1500 rpm, MAP for 20% boost
(NIMEPMBT 1.2 times unboosted NIMEPMBT).
166
is
35
-
-
I-
30
-
- -
-
.. .. .. . +
.
---
-
-
-
14
~-
25 -
-- *-PRF110
CP
20
-
-13
15 -
0
10
- - - - - - - - - - - -
-
PRF1 20
C.R.
PRF115
C.R.
- - -
- -~
-12
-
j
--
C0
0.
--
z
-
C.R.
-PRF105
C.R.
PRFI 00
C.R.
- - PRF120
NIMEP
PRFII5
- -- NIMEP
PRF115
-
--
-
--
-
--
50
-11
-PRFI10
0-
NIMEP
-5
i
9
10
12
11
1
13
410
PRF105
NIMEP
14
---
PRFIOO
NIMEP
Rc
Figure E. 7 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for PRFs; X = 1.0, 1500 rpm, MAP for 40% boost
(NIMEPMBT 1.4 times unboosted NIMEPMBT).
16
35
-M--TRF105
C.R.
I-
30
-
15
IrOO000000000000-000a
O
25
-
20
-
14 'C'
CD
10
C.R.
-- +-TRF95
C.R.
- -000-0-
15
-- 4-TRF100
.
-
-- w-TR90
C.R.
-
-TRF105
)-
NIMEP
13 z
-
-TRFIOO
- --
0
c
NIMEP
5-
E
'0
12
0-5
- -~-
NIMEP
-
'
9
1
10
I1
11
1
13
12
13
11
TRF95
- TRF90
NIMEP
14
Rc
Figure E. 8 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased re for TRFs; X = 1.0, 1500 rpm, MAP for 40% boost
(NIMEPMBT 1.4 times unboosted NIMEPMBT).
167
35
30
13
-
-
--12
0a
25
-
~---PRF1 05
20 -
C.R.
-- h- PRFI00
C.R.
-- +-PRF95
C.R.
- -- PRF115
11
CL
I
.0
E
0
C.
PRF115
C.R.
PRF110
C.R.
15
-
10
-
-- - -- - -- - --
- -- - --
-
102
NIMEP
5 -
NIMEP
9
0 ---5 i
9
'PRF110
- --
- -x- PRF105
-
NIMEP
I
I
1
1
.
10
11
12
13
14
-
-PRFlOO
-
-
8
NIMEP
PRF95
NIMEP
Rc
Figure E. 9 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for PRFs; X = 1.3, 1500 rpm, MAP for 15% boost
(NIMEPMBT 1.15 times NIMEPMBT at stoichiometric unboosted WOT).
35
30
T3
-13
-
-
-
-.-
-
-
25
-
20
-
15
-
-
1-
E
0
.
- - - - - - - -
- - - - -
12
--
-- Ar--TRF100
C.R.
-- +-TRF95
C.R.
-112
-.-
TRF90
C.R.
-
10
TRF105
C.R.
-
-
NIMEP
-
)I(-
0
-TRF105
--
10 i
A-
NIMEP
5-
9
-
*-
-
-
8
-5
10
12
11
13
TRF95
NIMEP
01+--
9
-TRF1OO
-TRF90
NIMEP
14
Rc
Figure E. 10 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased re for TRFs; X = 1.3, 1500 rpm, MAP for 15% boost
(NIMEPMBT 1.15 times NIMEPMBT at stoichiometric unboosted WOT).
168
35
15
-
----
30
14
25
-
20
-
PRF1 20
C.R.
PRF15
C.R.
-- *-PRF110
C.R.
-- X- PRF1 05
13j
C.R.
0
15 -
.
. . .
-
-
:
- -
+
E
0
10
12
-
---
PRFIOO
- -+--
C.R.
PRF120
NIMEP
'--x-
- --
5 -
NIMEP
11
C)
PRF115
-
<- -PRF110
-
-x- PRF105
-
e-
0-
NIMEP
i 10
-5 i
9
10
12
11
13
NIMEP
14
-PRFIOO
NIMEP
Rc
Figure E. 11 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for PRFs; X = 1.3, 1500 rpm, MAP for 30% boost
(NIMEPMBT 1.3 times NIMEPMBT at stoichiometric unboosted WOT).
.
35
30
- 15
-
-
-- '-TRF100
-
14
C.R.
25 I-
20
---+-TRF95
C.R.
-
-
13 '
0
U
..
15
-
10
-
-U-TRF90
C.R.
-
E
C
I-TRF105
C.R.
<- -TRF105
NIMEP
-12 2
- --
-TRF1OO
NIMEP
5-
11
0-5
' 10
i
9
10
12
11
13
-
--
-
-
TRF95
NIMEP
-TRF90
NIMEP
14
Rc
Figure E. 12 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for TRFs; ) = 1.3, 1500 rpm, MAP for 30% boost
(NIMEPMBT 1.3 times NIMEPMBT at stoichiometric unboosted WOT).
169
35
0
30
41
2
16
---
PRF110
C.R.
-- - -
25
- -
RF0
C.R.
- 15
-
PRF100
C.R.
E
2
4
4 ,
-----------
~15
C
-
---
W
-
1
-
---
-x- -PRF105
NIMEP
-
12
0 -- - - - - - -
0
4-PRFIOO
-
- - -
- - - - - - - - - - -NME
- --
511
9
PRF110
NIMEP
,
E
PRF95
C.R.
PRF90
C.R.
--
---------
-PRF95
NIMEP
10
11
12
13
14
-.a--PRF90
NIMEP
Rc
Figure E. 13 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for PRFs; X = 1.0, 1500 rpm, MAP for 40% boost
(NIMEPMBT 1.4 times unboosted NIMEPMBT), 15% fuel reformate fraction.
16
35
0 30 -R
30-
-416TRF105
C.R.
-- *--TRF100
9C
. . . . -. .
25
- 15
C.R.
--
S20 -
-
TRF95
C.R.
14 V
-- m-TRF90
U
13 z
-
C.R.
10-
-
-- 12
E
S0---
-
-
11
9
10
12
11
13
- -
--
TRFI00
NIMEP
TRF95
NIMEP
---------------
-5
-TRF105
NIMEP
TRF90
NIMEP
14
Rc
Figure E. 14 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased re for TRFs; X = 1.0, 1500 rpm, MAP for 40% boost
(NIMEPMBT 1.4 times unboosted NIMEPMBT), 15% fuel reformate fraction.
170
16
35
PRF100
C.R.
+- PRF95
C.R.
-+-
30 15
P
-
25-
--
0
--
20
.a
'
-
- --
15
-
10
-
- - -4"-
..-
""~
--
-- m-PRF90
o
-U---
14
-
cc
~~~~
0
13 z
NIMEP
512
LI)
E
0
C.R.
PRF85
-+C.R.
- PRF80
C.R.
- -- -PRFIOO
-
+-. PRF95
-
NIMEP
-- -PRF90
NIMEP
-
-PRF85
011
-5
9
12
11
10
13
NIMEP
14
--
-PRF80
NIMEP
Rc
Figure E. 15 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for PRFs; X = 1.0, 1500 rpm, MAP for 40% boost
(NIMEPMBT 1.4 times unboosted NIMEPMBT), 30% fuel reformate fraction.
-16
35
L-
30
-
15
25 -
-4(-TRF105
C.R.
-- *-TRF100
C.R.
-- +-TRF95
C.R.
----
20 -
14 V.
a.
. - - --
0
15 -
TRF90
C.R.
-4--TRF85
C.R.
-+-TRF80
C.R.
-13 z
10 -
-
-x- TRF105
NIMEP
-TRF100
- -(U
E
0
NIMEP
5-
0-5
-- - -
- - - - - - - - - - - - - - - - - -
' 11
'
9
12
10
12
11
Rc
13
14
-
-
TRF95
NIMEP
- .0-
TRF90
NIMEP
+-o TRF85
NIMEP
---
TRF80
NIMEP
Figure E. 16 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for TRFs; X = 1.0, 1500 rpm, MAP for 40% boost
(NIMEPMBT 1.4 times unboosted NIMEPMBT), 30% fuel reformate fraction.
171
,
35
I-
'-
a)
11
-r-kPRFI0O
C.R.
30 25
-
20
-
15
-
10
-
- - - - - -- - .
.- - - - - -
,
0
I
,
-PRF95
.
- 10
C.R.
-U-PRF90
C.R.
-+-PRF85
-9
C.R.
PRF80
C.R.
-
CL
-
.'i
.PRFI0O
NIMEP
0
-
5E
0
PRF95
+-
NIMEP
-7
-PRF90
-
0-4---
NIMEP
-
-5 t
1
9
10
a
11
12
14
13
- -+-
PRF85
NIMEP
PRF80
NIMEP
Rc
Figure E. 17 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased re for PRFs; X = 1.3, 1500 rpm, MAP set to match
NIMEPMBT at stoichiometric unboosted WOT, 15% fuel reformate fraction.
11
35
-
-----
30
-- *-TRF100
-
- - --
C.R.
-+-TRF95
C.R.
-
10
I-
25 -
--
TRF90
C.R.
0
20
-4-TRF85
'U
15 10
8 2
-
--
5 7
0 -- -
-
TRF95
NIMEP
-
e-
-
-0-
TRF90
NIMEP
6
-5
9
C.R.
'TRF1OO
- --
NIMEP
*0
E
0
C.R.
-+--TRF80
10
12
11
Rc
13
14
TRF85
NIMEP
- --
-TRF80
NIMEP
Figure E. 18 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased re for TRFs; X = 1.3, 1500 rpm, MAP set to match
NIMEPMBT at stoichiometric unboosted WOT, 15% fuel reformate fraction.
172
35
11
..
-
I-
30
-.-.
-E-PRF90
- - - --..
C.R.
-
- -- -
- -..
- -
Ca
--
10
25 20
-
15
-
----
--
-
-
-
-
-
-
-.)--
--
-
-+-
-
-
-
8 z
10 -
U
0
--
C.R.
PRF75
C.R.
PRF70
C.R.
PRF90
NIMEP
- -o- PRF85
5-
-- 7
--
E
0
-
PRF85
C.R.
PRF80
NIMEP
- -+-
0 -
NIMEP
-
6
-5
9
PRF80
10
NIMEP
14
13
12
11
-K- -PRF75
- --
PRF70
NIMEP
Rc
Figure E. 19 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for PRFs; X = 1.3, 1500 rpm, MAP set to match
NIMEPMBT at stoichiometric unboosted WOT, 30% fuel reformate fraction.
11
35.- 4----
00a
0
30
-
.
-.
--...-
-
---
-U-TRF90
.
,------
--
---
e
10
25
-
20
-
15
-
-- +-TRF80
9':
.
- - - - - --- -- - - ---
-
10-
8 z
C.R.
---- TRF75
C.R.
TRF70
C.R.
- -TRF90
- --
57
0
E
0
C.R.
TRF85
C.R.
NIMEP
TRF85
NIMEP
- -~- -TRF8O
0-
NIMEP
- -K-
-.
9
10
12
11
Rc
13
13
6
14
TRF75
NIMEP
- --
TRF70
NIMEP
Figure E. 20 - Increase of combustion retard to just avoid knock and change of
NIMEP with increased r, for TRFs; X = 1.3, 1500 rpm, MAP set to match
NIMEPMBT at stoichiometric unboosted WOT, 30% fuel reformate fraction.
173