Liquid Fuel Hydrocarbon Emissions Mechanisms in
Spark-Ignition Engines
by
Gary B. Landsberg
B.S., Mechanical Engineering
University of South Alabama, 1998
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 Institute of Technology
June 2000
@ 2000 Massachusetts Institute of Technology
All rights reserved
Signature of Author
Department 6f Mechanical EngTneering
May 1, 2000
/
Certified by
Jokn B. Heywood
Sun Jae Professor of Mechanical Engineering
Thesis Supervisor
Accepted by
Ain A. Sonin
Chairman, Department Committee on Graduate Students
MASSACHUSETTS INSTITUTE
OF TECHNOLOGY
SEP 2 0 2000
LIBRARIES
Liquid Fuel Hydrocarbon Emissions Mechanisms in
Spark-Ignition Engines
by
Gary B. Landsberg
Submitted to the Department of Mechanical Engineering
May 2000 in Partial Fulfillment of the Requirements
for the Degree of Master of Science in Mechanical Engineering
ABSTRACT
Future regulations, such as the ultra low emissions vehicle (ULEV) specified by the
California Air Resources Board (CARB), are placing stricter limits on the allowable hydrocarbon
(HC) emissions and make the understanding of the sources of HC emissions of considerable
importance. Previous researchers have identified six sources of hydrocarbon emissions: oil
layers, deposits, liquid fuel, quenching, crevices, and exhaust valve leakage. The objective of
this research was to gain a better understanding of the liquid fuel contribution to hydrocarbon
emissions.
The purpose of this work was to develop a fundamentally based description of liquid fuel
transport into the engine cylinder of a port fuel injected, gasoline fueled, SI engine, and to
develop a method of quantifying the liquid fuel contribution to HC emissions. To simulate the
liquid fuel flow from the valve seat region into the cylinder, a specially designed injector probe
was constructed and used to inject controlled amounts of liquid fuel onto the port wall close to
the valve seat. By fueling the engine on a gaseous fuel, either propane or pre-vaporized
Indolene, and injecting liquid fuel close to the valve seat, we can examine the effects of liquid
fuel entering the cylinder at different circumferential locations around the valve seat. These
experiments were carried out with both open and closed valve fuel injection to assess the
differences that residual blowback and evaporation produce.
The location of liquid fuel around the valve seat was found to have a significant impact
on engine-out hydrocarbon emissions. These experiments indicated that for all amounts of liquid
fuel injected at the valve seat, the fuel delivered closest to the exhaust valve resulted in the
highest engine-out HC's, while the location farthest from the exhaust valve had the lowest HC's.
The differences between closed valve and open valve injection showed similar trends for all
probe locations, indicating that the blowback during valve overlap has the same impact on
mixture preparation and vaporization at all locations. Comparison with other experimental data
suggest that the location of liquid fuel in the cylinder after induction into the engine may be a
major source of the differences in HC's between the different probe locations around the intake
valve.
Thesis Advisor: Professor John B. Heywood
Title: Sun Jae Professor of Mechanical Engineering
3
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4
Acknowledgments
I guess my journey here is over, and as with every journey there were high points and low points along the way.
I've done many new things, learned new ideas, and met lots of great people along the way. I would first and foremost
like to thank my tour guide (pronounced "advisor"), Professor John B. Heywood, he kept me from getting lost along
the way and taught me many of the things I needed to survive. I would also like to thank Professor Wai Cheng for
always taking the time to help me fix my electronic gremlins. I could never have gotten my project going without the
help of Brian Corkum and Matt Rublewski, I enjoyed working with you both. Brian, thanks for showing Matt how to
check to see if a spark plug is sparking. Matt, thanks for showing me what happens when you become the ground for a
sparking spark plug (zap zap zap).
Of all the people I know, there is only one person who truly understands what I have gone through in the past
two years and that is Brian Hallgren (aka. Green Halogen). As my officemate, friend, mountain bike partner, and coconspirator, we have had an adventure nearly everyday, bringing a sense of fun to our otherwise dull days at work. I
wish you best of luck on your Ph.D., don't let MIT get to you. One more thing, remember our motto, "it wasn't me, I
don't know where it is, you should order one."
I would also like to thank everyone in the Sloan Lab past and present. Special thanks to the Engine and Fuels
Consortium members for funding this research and providing valuable feedback at our meetings. The Consortium
members include: DiamlerChrysler, Ford, General Motors, ExxonMobil, Shell UK, and Volvo Car.
There is one person above all others, who helped me through my time here and that is my best friend and wife,
Maria. Throughout the past nine years you have given me the courage and confidence that has allowed me to get to
where I am. Thank you Maria, for helping me to get through this with my sanity and sense of humor. Coming home to
you was (and is) always the best part of my day. I would also like to sincerely thank Annabelle, you have always
listened to me, calmed me down after losing my temper, and have stuck by my side everyday for the past six months.
I could have never done any of this without the help of my entire family. I would like to thank my parents for
always supporting me in all of my endeavors. Brad, thanks for letting me live vicariously through your adventures
during my term here. For the next two years when you are in grad school, I will do the same for you.
Last but not least I would like to thank all of the people who have helped to get me to MIT. Including the entire
engineering faculty at the University of South Alabama, most especially Professors Eugene Odell, Jay Kapat (now at
UCF), Lanier Cauley, Ali Engin, and David Hayhurst.
When I was about three years old I took a diaper pin stuck it into an electrical socket and
discovered electricity. After my mom unplugged me, I thought, "Ouch, I'll never ever do that
again ". MIT was a similar experience - I learned a lot of things I'll neverforget. -GL
5
TABLE OF CONTENTS
Abstract
Acknowledgments
Nomenclature
List of Tables
List of Figures
3
5
8
9
10
CHAPTER 1 INTRODUCTION .....................................................................................
15
1.1
1.2
1.3
1.4
1.5
1.6
1.7
Mixture Preparation in SI Engines .............................................................
Hydrocarbon Emissions in the Federal Test Procedure ................................
Sources of Hydrocarbon Emissions.............................................................
Previous Work on Liquid Fuel Hydrocarbon Emissions................................
Project Scope .............................................................................................
Tables.............................................................................................................
Figures ...........................................................................................................
CHAPTER 2 EXPERIM ENTAL METHOD ..................................................................
2.1
2.2
2.3
2.4
2.5
2.6
2.7
2.8
2.9
Nissan Test Engine ......................................................................................
Nissan Engine Control System ...................................................................
Engine M easurement System .......................................................................
In-Cylinder Pressure M easurement...............................................................
Hydrocarbon M easurement System.............................................................
Labview Data Acquisition Program.............................................................
Engine Operating Conditions .......................................................................
Tables.............................................................................................................
Figures ...........................................................................................................
15
16
16
17
21
23
25
35
35
35
36
37
37
38
38
41
45
CHAPTER 3 EFFECTS OF ENGINE OPERATING PARAMETERS ON HC.............. 49
3.1
3.2
3.3
3.4
3.5
3.6
3.7
3.8
Factors Affecting Liquid Fuel Distribution .................................................
Intake Airflow Processes .............................................................................
Effects of Spark Timing ..............................................................................
Effects of Relative Air/Fuel, Lambda ..........................................................
Coolant Temperature Experiments...............................................................
Propane and Indolene Fueling .....................................................................
Effects of Fuel Injection Timing ...................................................................
Figures ...........................................................................................................
6
49
49
50
51
52
52
53
55
CHAPTER 4 LIQUID FUEL INJECTION PROBE EXPERIMENTS ............................
4.1
4.2
4.3
4.4
4.5
4.6
4.7
4.8
4.9
4.10
4.11
Overview of Liquid Fuel Injector Probe Experiments ...................................
Pre-V aporizing Injector ..............................................................................
Liquid Fuel Injector Probe ..........................................................................
Liquid Fuel Injector Injection Timing Sweeps .............................................
Fuel Probe Indolene/Propane Fuel Sweeps..................................................
Fuel Probe Fuel Sweeps ..............................................................................
Comparison of OVI and CVI with Injector Probes.......................................
Experim ental Fuel Probe Comparison with PFI...........................................
Discussion ..................................................................................................
Tables.............................................................................................................79
Figures ...........................................................................................................
CHAPTER 5 SUMMARY AND CONCLUSIONS ..........................................................
5.1
5.2
5.3
5.4
Liquid Fuel Transport into Cylinder ..............................................................
Comparison of Liquid Fuel Probes to Standard PFI.......................................
Conclusions..................................................................................................
Figures .........................................................................................................
REFEREN CES ................................................................................................................
7
71
71
71
72
73
74
74
76
77
78
81
105
105
106
106
111
115
NOMENCLATURE
BDCC
- bottom dead center of compression stroke
CFI
CVI
ECU
FID
FTP
HC
MAF
MAP
-
N-IMEP
NOX
OVI
PFI
- net indicated mean effective pressure
- oxides of nitrogen
- open valve injection
- port fuel injection
SMD
TDC
- sauter mean diameter
- top dead center
TDCC
- top dead center of compression stroke
central fuel injection
closed valve injection
engine control unit
flame ionization detector
federal test procedure
hydrocarbons
mass air flow
manifold air pressure
8
LIST OF TABLES
Table 1.1
Table 1.2
Table
Table
Table
Table
2.1
2.2
2.3
2.4
Table 4.1
Table 4.2
Engine test matrix used by Stanglmaier [14]...............................................
Summary of experimental data on HC emissions during startup, effects of liquid fuel in the cylinder........................................................
23
N issan engine specifications .........................................................................
National Instruments data acquisition specifications....................................
Data output variables from data acquisition program....................................
Conversion factors for raw voltage input to data acquisition
..
sy stem ..........................................................................................................
41
41
43
HC fuel fraction values based on linear regression analysis of
experimental data with different assumed Z values......................................
HC fuel fraction values for standard engine operating
con ditions.....................................................................................................
9
23
43
79
. 79
LIST OF FIGURES
Figure
Figure
Figure
Figure
Figure
1.1
1.2
1.3
1.4
1.5
Figure 1.6
Figure 1.7
Figure 1.8
Figure 1.9
Figure 2.1
Figure 2.2
Figure 2.3
Figure 2.4
Federal test procedure, FTP 75 .....................................................................
Steady-state HC emissions flowchart [4]......................................................
Crevices in the combustion chamber [2]........................................................
Effect of crevice volume during warm-up [6]...............................................
Plot of total, engine-out HC emissions versus time after coldstart for two standard injection (PFI) and two pre-vaporized
injection (PV -CFI) starts [18]........................................................................
Liquid fuel distribution in the port and cylinder during coldstart and warm engine operation [17] ............................................................
Schematic of liquid fuel injector probe used by Stanglmaier
[14 ] ....................................................................................................................
Liquid fuel injection location for direct injection of fuel used
by Stanglm aier [14][15] ................................................................................
Effect of wall wetting location on engine-out HC's,
Stanglm aier [14]............................................................................................
25
25
27
27
29
29
31
31
33
Standard intake runner and intake port geometry for the Nissan
. 45
Sentra engine................................................................................................
45
location
.................................
pressure
transducer
Top-view of in-cylinder
47
Side-view of pressure transducer location [24].............................................
47
Pressure transducer calibration data...............................................................
Figure 3.1
Computed intake backflow distance and volume, 1200 rpm,
cylinder displacement volume is 486cc[26]...................................................
Figure 3.2 Calculated mass flow rates through intake valve at 0.4, 0.6, and
0.85 B ar MA P, [27] .......................................................................................
Figure 3.3 Intake and exhaust stroke in-cylinder pressure data .....................................
Figure 3.4 Cylinder pressure during intake and exhaust .................................................
Figure 3.5 Spark timing vs. HC with matched N-IMEP for propane and
..
In d o len e........................................................................................................
Figure 3.6 Maximum cylinder pressure vs. spark timing with matching of
N-IMEP for propane and Indolene Mixtures .................................................
Figure 3.7 Spark timing vs. HC with Nissan injector and propane ................................
Figure 3.8 Nissan Injector and Propane fueled N-IMEP vs. spark timing,
w ith M A P = 0.5 bar .......................................................................................
Figure 3.9 Nissan Injector and Propane fueled N-IMEP vs. location of
peak cylinder pressure, with MAP = 0.5 bar.................................................
Figure 3.10 Location of peak cylinder pressure vs. spark timing .....................................
Figure 3.11 HC vs. Lambda for propane and pre-vaporized Indolene
..
experim ents ................................................................................................
Figure 3.12 N-IMEP vs. Lambda for propane and pre-vaporized Indolene......................
10
55
55
57
57
59
59
61
61
63
63
65
65
Figure 3.13 Effects of Lambda on cyclic variability ........................................................
Figure 3.14 HC engine warm-up experiments for propane and standard
N issan injection of Indolene ..........................................................................
Figure 3.18 Liquid fuel sweeps with Nissan and BMW injectors, OVI ..............
Figure 3.19 Injection timing vs. HC for standard Nissan Injector and the
low-flow BMW injector. Injection timing corrected for 175
degree transport delay ....................................................................................
Figure
Figure
Figure
Figure
4.1
4.2
4.3
4.4
Figure 4.5
Figure 4.6
Figure 4.7
Figure 4.8
Figure 4.9
Figure 4.10
Figure 4.11
Figure 4.12
Figure 4.13
Figure 4.14
Figure 4.15
Figure
Figure
Figure
Figure
Figure
Figure
Figure
4.16
4.17
4.18
4.19
4.20
4.21
4.22
67
67
69
69
81
Schematic of pre-vaporizing injector.............................................................
81
Air-Assist mass airflow estimation...............................................................
83
Liquid fuel injector probe locations ...............................................................
Liquid fuel injection delay through fuel probes, based on liquid
83
fuel probe visualization study ........................................................................
Probe B Liquid fuel probe injection timing sweep, effect of fuel
injection timing on HC and N-IMEP, 36% Indolene, balance
. 85
P rop ane .........................................................................................................
effect
of
fuel
sweep,
Probe B Liquid fuel probe injection timing
injection timing on HC and Cyclic Variation of N-IMEP
85
(Covariance), 36% Indolene, balance Propane ............................................
Probe A Liquid fuel probe injection timing sweep, effect of
fuel injection timing on HC and N-IMEP, 36% Indolene,
87
balance Propane ............................................................................................
Probe A Liquid fuel probe injection timing sweep, effect of
fuel injection timing on HC and Cyclic Variation of N-IMEP
87
(Covariance) , 36% Indolene, balance Propane ............................................
Probe C Liquid fuel probe injection timing sweep, effect of fuel
injection timing on HC and N-IMEP, 36% Indolene, balance
. 89
P rop an e .........................................................................................................
of
fuel
sweep,
effect
timing
Probe C Liquid fuel probe injection
injection timing on HC and Cyclic Variation of N-IMEP ............................. 89
Comparison of HC's for probes A,B, and C. 36% Indolene,
91
balance Propane ............................................................................................
Comparison of Covariance for probes A,B, and C. 36%
91
Indolene, balance Propane ............................................................................
93
Indolene/propane fuel sweep CVI.................................................................
93
Indolene/propane fuel sweep OVI ................................................................
Comparison of standard deviations of N-IMEP for liquid fuel
. 95
probe sw eep ..................................................................................................
Comparison of OVI and CVI HC's for probes A and C............................... 95
97
Liquid fuel probe sweep CVI HC's ..............................................................
97
Liquid fuel probe sweep OVI HC's ..............................................................
99
Effect of liquid fuel on exhaust HC's, CVI ...................................................
99
Effect of liquid fuel on exhaust HC's, OVI ...................................................
101
Effect of liquid fuel probe mass on N-IMEP, CVI .........................................
101
Effect of liquid fuel probe mass on N-IMEP, OVI .........................................
11
Figure 4.23 Correlation of N-IMEP with the standard deviation of N-IMEP
for all liquid fuel probe sweep data................................................................
Figure 5.1
Figure 5.2
Figure 5.3.
CFD Simulation of 4-valve head, at WOT with one intake
valve closed, 1500 rpm.[28]
Nissan head configuration
o v erlay .............................................................................................................
Relative liquid fuel transport mechanisms for probe locations
"A " and "C " ....................................................................................................
Relative comparison of liquid fuel transport mechanisms for
different probe locations ...............................................................................
12
103
111
111
113
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13
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14
CHAPTER 1
INTRODUCTION
1.1 Mixture Preparation in SI Engines
Spark Ignition engines are traditionally fueled by liquid fuel (gasoline) vaporizing and
being entrained into the airflow entering the engine cylinder. The quality and repeatability of the
combustion process in the engine combustion chamber is a function of the fuel-air mixture
homogeneity and the mixture fuel-air ratio. This mixing of the fuel and air to produce a burnable
mixture is referred to as mixture preparation.
Early spark ignition engines were fuel by carburetors where liquid fuel is injected into the
intake system at a central location well upstream of the intake valves. The carburetor fuel flow is
metered by a venturi tube, which changes the fuel flow depending on intake airflow. More precise
fuel control was achieved with the advent of the electronic fuel. Initially, fuel injectors injected
fuel upstream of the throttle into the airflow, this type of electronic fuel injection is known as
single point injection or central fuel injection (CFI).
In more modern engines, each engine
cylinder has its own fuel injector and liquid fuel is injected directly into the intake port of each
cylinder; this is referred to as multipoint fuel injection or port fuel injection (PFI). Electronic fuel
injection metering is controlled using an engine control unit (ECU) which injects the appropriate
fuel amount based one or more of the following: mass airflow sensors, manifold intake pressure
sensors located in the intake port, throttle position, or a lambda sensor in the exhaust manifold. A
more detailed description of these fuel metering systems can be found in [1] and [2].
Regardless of the type of fuel metering system, the liquid fuel must become mixed with the
air and evenly distributed into the engine cylinders. Liquid fuel normally enters the airflow in the
form of drops. These liquid fuel drops though partially vaporized are deposited onto the port wall
and valve forming a fuel film, or transported directly into the engine cylinder. These transport
processes are highly dependent on engine temperature, fuel injection location, intake port
geometry, fuel injector spray pattern, engine speed, engine load, and fuel properties. Under ideal
conditions, all the liquid fuel should be vaporized and mixed uniformly with the air, as the mixture
enters the engine cylinder. However, real engines and real operating conditions are not ideal.
Previous work has shown that with port fuel injection liquid fuel is deposited on the intake valve
and surrounding intake port walls for all operating conditions [3].
15
1.2 Hydrocarbon emissions in the Federal Test Procedure
Federal Regulations require that all passenger vehicle emissions be measured using a
standardized test which represents standard driving conditions. In the United States this test is
called the Federal Test Procedure or FTP 75. A typical FTP driving cycle for a 90's model Saturn
1.9 L, is shown in Figure 1.1.
The first cycle of the FTP (120 seconds) simulates the cold-start and warm-up period of
engine operation. During this period, prior to catalyst light-off, tail-pipe HC emissions are high.
Several factors contributing to the extra HC emissions are larger crevice volumes, higher gas
density in crevices, increased oil absorption, lower exhaust port oxidation, larger quench volume,
and substantial liquid fuel on the cylinder walls and in crevices. Hydrocarbon emissions during
the first cycle (Bag 1) of the FTP emissions test are the source of 74-94% of the total emissions for
FTP test.
Previous studies have indicated that liquid fuel in the cylinder may be a major
contributor to HC emissions during cold operation.
Future regulations, such as the ultra low
emissions vehicle (ULEV) specified by the California Air Resources Board (CARB), are placing
stricter limits on the allowable HC emissions.
Future emissions regulations make the
understanding of the sources of HC emissions during cold-start and warm-up of considerable
importance.
1.3 Sources of Hydrocarbon Emissions
The six sources of the unburned hydrocarbons: oil layers, deposits, liquid fuel, quenching,
crevices, and exhaust valve leakage are the primary focus of the following sections.
discussion of liquid fuel hydrocarbons is contained in section 1.4.
The
Cheng [4] developed a
flowchart which outlines the unburned hydrocarbon emissions mechanisms in a spark-ignition
engine during steady-state engine operation, this flowchart is shown in Figure 1.2.
Currently, the largest source of HC is believed to be crevices in the combustion chamber.
Crevices are narrow regions connected to the combustion chamber where unburned gas can escape
combustion. Previous research indicates that during steady-state engine operation approximately
5% of the injected fuel escapes initial combustion due to crevices [4].
The crevices in the
combustion chamber where HC may escape combustion are the cylinder head gasket crevice, spark
plug crevice, valve seat crevice, piston ring pack crevice, and the piston top land crevice as shown
16
in Figure 1.3. The piston top land crevice is the volume between the cylinder wall and piston
extending from the piston crown to the top piston compression ring. The piston ring pack crevice
is the volume contained in the piston ring grooves and clearances and is believed to account for
approximately 85% of the total crevice volume according to Min [5]. Sterlepper conducted several
experiments to study the HC emissions during cold-start and warm-up using three different pistons
with different crevice volumes [6]. The engine warm-up data shown by Sterlepper's data in Figure
1.4 indicates that the HC emissions during cold-start are virtually independent of crevice volume.
However, as the engine warms-up the HC emissions decrease to different steady-state values.
Sterlepper states that at start up, the cool wall temperatures cause the flame to quench before
reaching the piston crevices.
Recent experimental studies on deposits suggest that during steady-state engine operating
conditions between 0.2 and 0.5% of the injected fuel was emitted as unburned HC due to deposits
[7]. Previously, Cheng et al. estimated this value to be about 1% [4].
Under warm operating conditions the oil layer mechanism accounts for less than 10% of
the total HC emissions [8]. Research by Hochgreb and Kaiser suggest that approximately 80% of
the desorbed fuel is oxidized when returned to the cylinder [9],[10]. Due to lower temperatures
and a corresponding higher absorption of gasoline into the oil layer, the oil layer contribution to
HC emissions is expected to be higher at lower temperatures [4][8]. Initially a large amount of
fuel is absorbed into the oil reaching a maximum value 50 seconds after cold start [11].
Cheng et al. [4] estimated that under steady-state operation 0.5% of the injected fuel is
emitted as unburned HC due to flame quenching at the cylinder walls.
During cold engine
operation the wall temperatures are cooler and the quench layer is expected to increase in thickness
leading to an increase in HC.
Experimental studies indicate that 5-7% of the engine out Hydrocarbon emissions may be
due to exhaust valve leakage [12][13]. These leaks are likely due to valve and seat distortion due
to thermal and other loads; and by deposit flakes being trapped in the valve seat [13].
1.4 Previous Work on Liquid Fuel Hydrocarbon Emissions
Previously, Cheng et al. [4] estimated that 1.2% of the injected fuel is emitted as
unburned HC due to liquid fuel in the cylinder. Several researchers have studied the effects of
17
liquid
fuel
during
steady-state
engine
warm
operation
and
during
warm-up
[14][15][16][17][18][19].
Min and Kaiser conducted a series of experiments using a pre-vaporizing injector during
engine warm-up. Min et al. [16] measured the engine out hydrocarbon emissions for an engine run
on pre-vaporized Indolene and normal liquid injection of Indolene. Min found that the HC levels
were four times higher during the first 50 seconds of engine operation when using standard port
fuel injection compared to pre-vaporized injection.
Kaiser et al. [18] conducted an experiment similar to Min's where standard port fuel
injection (PFI) and pre-vaporized central fuel injection (PV-CFI) were compared during a 23 'C
cold start. In Kaisers' study, special care was taken to inject the same amount of fuel for both
injection methods and fuel-rich operation was avoided (this minimizes the effect of fuel control
and allows for comparison)[18].
A surprising result of this study was that under light load
conditions the fuel preparation can have negligible effect on the engine-out HC as can be seen in
Figure 1.5.
Imatake and Kudo [17] developed an engine experiment where the intake and exhaust
valves are electronically sealed at a specific point during firing operation. The intake port and
cylinder are sealed and both the cylinder and port are vented with 200'C air (200'C is the end
point of distillation for the given test fuel). This method allows for total vaporization of the fuel,
the vented air and vaporized fuel are then sampled with a flame ionization detector (FIID). These
researchers found that during warm-up 9 times the injected amount of fuel is stored as a fuel film
on the intake port wall (135 mg on port, standard injection is 15 mg per cycle). The corresponding
in-cylinder wall wetting was 2.5 times the injected amount. Under steady-state warm conditions
(80'C coolant) 3 times the injected amount is stored on the intake port and 0.8 times the injected
amount is found in-cylinder wall wetting. Imatake found that, during warm-up, the cylinder
contained 37mg of liquid fuel; and 12mg during steady-state operation. The normal injection
amount for this study was 15 mg per cylinder. The wall wetting amount in the port and cylinder
for cold-start and warm engine operating conditions are shown in Figure 1.6. This same technique
was used to study the effects of open valve injection (OVI) vs. closed valve injection (CVI). They
found that with open valve injection 18% more of the injected fuel is transported directly into the
cylinder resulting in 16% more in-cylinder wall wetting. They also found that with open valve
18
injection there is 7% less port wall wetting but this port wall film has a 12% lower evaporation
rate than the port wall film for closed valve injection.
Shayler et al. [19] performed a fuel audit on an unmodified Ford Zetec 1.81, 16 valve, four
cylinder engine during cold-start. Shayler found that for a 20'C cold start 20 mg and 10mg of
liquid fuel are stored in the cylinder during the first 50 and 100 seconds of engine operation
(normal injection amount was 20 mg/injection).
Fry and Nightingale [20] found that under certain operating conditions a wall film was seen
to form in the crevices around the inlet and exhaust valves and the apex of the pent-roof head.
This wall film would form during induction and remain unburned throughout the combustion
process. This film would then be sucked out of the cylinder during the exhaust blowdown process
as liquid fuel. This tendency to form a wall film in this location was stronger for CVI than for
OVI for all load and speed conditions studied.
Witze and Green at Sandia conducted a visualization study using LIF and Flame-Emission
imaging of liquid fuel films and pool fires [21]. The engine setup consisted of a Bowditch piston
with a window in the piston crown to provide optical access to 80% of the combustion chamber.
The engine head was a standard DOHC with four valves per cylinder, mounted on a single cylinder
crankcase. At all engine coolant temperatures 20, 40, and 60 'C there was a significant difference
in the distribution of the pool fires for CVI and OVI. CVI generally resulted in larger pool fires
below the intake valve squish region, possibly on the piston in some cases. OVI resulted in a
wider distribution of pool fires throughout the cylinder but primarily resulted in pool fires between
the exhaust valves. This pool fire between the exhaust valves suggest that it is likely that liquid
fuel is swept out of the cylinder during the exhaust stroke.
Stanglmaier studied the effect of in-cylinder wall wetting location on HC emissions using a
specially designed spark plug mounted injector probe [14][15]. This directional probe allowed
liquid fuel to be injected onto specific locations within the combustion chamber. This method
allows for the quantification of the fate of specific amounts of liquid fuel deposited at different
locations within the combustion chamber. This experimental probe is shown in Figure 1.7. The
fuel is deposited on the cylinder liner just below the head. This engine test rig set-up consisted of
an optically accessible engine that is geometrically identical to a GM Quad-4, 2.26L, 4-valve, 4cylinder engine. The engine was fueled on liquefied petroleum gas (LPG), which is approximately
95% propane, and a small amount of California Phase 2 fuel which was injected through the
19
injector probe. The engine was run slightly lean at X = 1.1, approximately 85% of the fuel was
LPG and approximately 15% of the fuel was liquid California Phase 2.
These experiments indicated that the location of the liquid fuel within the cylinder has a
very significant effect on the engine-out HC. The probe wall wetting locations are shown in
Figure 1.8 and the effects of wall wetting location on HC's is summarized in Figure 1.9. The
highest HC emissions were found to come from wall wetting at the E-E location, wall wetting on
the exhaust side of the cylinder in-between the exhaust valves. The next most significant HC
came from wetting of the piston top. The lowest HC came from wall wetting location I-I, wall
wetting on the cylinder in-between the intake valves. Stanglmaier suggest that the differences in
the HC for the different wall wetting locations can be explained in terms of the physical location of
the liquid fuel relative to the exhaust valves. He explains, liquid fuel deposited on the cylinder
wall is scraped into the piston top land during the compression stroke and then laid back on the
cylinder wall during expansion as a liquid fuel film. Any of this fuel film that survives the
combustion and post-flame oxidation process will be scraped into the roll-up vortex during the
exhaust stroke. The portions of the roll-up vortex closest to the exhaust valves have a higher
chance of be exhausted than the portions of the vortex farther from the exhaust valves [14][15].
Injection timing was varied from 90 degrees before exhaust TDC to 270 degrees ATDC exhaust
with both 90*C and for 36'C coolant temperatures. These experiments indicated that HC are only
weakly dependent on injection timing for all injection locations for both coolant temperatures.
Stanglmaier suggest that the liquid fuel evaporation from the combustion chamber surfaces is a
slow process relative to the engine cycle.
Yang and Kaiser conducted several engine experiments where the effects of fuel injection
droplet size and fuel injection timing were examined [22]. These experiments consisted of three
injectors: a production injector Sauter mean diameter (SMD) 300gm, an air-assisted injector SMD
40gm, and an air-forced injector SMD 14gm. They found that the droplet size had a large impact
on HC's during open valve injection at lower coolant temperatures.
At lower coolant
temperatures, the HC's increase by 40% and 80% for A/F = 12.7 and A/F = 16.2 respectively for
the standard injector (300 gm). The HC's did not change outside of the experimental error for the
air-forced injector (14gm) during open and closed valve injection at the lower coolant temperature
(30'C) and the overall HC level was lower for this injector compared to the standard injector.
20
During warm engine operation 89'C coolant the droplet size had a much lower impact on the
engine-out HC levels for both OVI and CVI during rich and lean fueling. The standard injector
showed no significant change in HC's during fuel rich operation for OVI and CVI. The HC's were
20% higher for OVI vs. CVI injection for fuel lean operation with the standard injector.
1.5 Project Scope
The effects of liquid fuel on engine-out HC emission is not well understood at this point.
Additionally, the effects of liquid fuel are believed to be of increased importance during cold-start
engine operation when the crevice volumes are larger, the quench layer is thicker, and the port and
cylinder wall temperatures are lower. The relationship between liquid fuel location in the port and
subsequent location in the cylinder on HC emissions is also unclear. The purpose of this work is
to develop a fundamentally based description of liquid fuel buildup and transport into the engine
and to develop a method of quantifying the liquid fuel contribution to HC emissions. A novel
liquid fuel injector probe was developed which deposits controlled amounts of liquid fuel at three
precise locations around the intake valve seat. By controlling the amount and location of the
injected liquid fuel and measuring the engine-out hydrocarbon (HC) emissions we can estimate the
liquid fuel contribution to these emissions. The various injector probe locations allow us to
estimate the fate of liquid fuel entering the cylinder from different locations in the port and
deposited onto different locations in the cylinder.
21
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22
Part-Load
1500 rpm
1000 rpm
1.1
1.1
M
BT
MBT
0.345 bar, MAP 2.62 bar, BMEP
11%
18%
Idle
Speed
Lam bda
S park
Load
%Liquid Fuel
Table 1.1 Engine test matrix used by Stanglmaier [14]
0-50 seconds
0-100 seconds
Min [16], WOT, 900 rpm, PV
4250 ppm C1
4100 ppm C1
Min [16], WOT, 900 rpm, PFI
12000 ppm C1
9750 ppm C1
Kaiser [18], 3.72bmep, 1200 rpm, PV
3250 ppm C1
3025 ppm C1
Kaiser [18], 3.72bmep, 1200 rpm, PFI
4060 ppm C1
3615 ppm C1
Kaiser [18], 0.74bmep, 1200 rpm, PV
2500 ppm C1
2350 ppm C1
Kaiser [18], 0.74bmep, 1200 rpm, PFI
2500 ppm C1
2350 ppm C1
Shayler [19], 0.7 MAP, 2000 rpm PFI
20 mg
10mg
Imatake (17], 1200 rpm, 40Nm, open valve inj. 37 mg, 30deg coolant
12 mg, 80 deg coolant
Table 1.2 Summary of experimental data on HC emissions during startup, effects of liquid fuel in the cylinder.
PV-pre-vaporized fuel, PFI- normal port fuel injection.
*- this data is for different steady-state coolant temperatures not for the
transient start-up conditions.
23
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24
.80
300~
iBag
275enf
0
U-
en
C-)
225200175 150
12510
2
a.
I
Bag
-
-20
--20
-40
I-.......
_0U
0
1000
500
2000
1500
Time (seconds)
Figure 1.1 Federal test procedure, FTP 75
1 Fuel (100%)
91.5%
8.5%
Flame converts fuel to
CO 2, CO, H2O, H2 etc.
,
HC Mechanisms
1'
Fuel Only
-
-- q. Fuel
SLi 1.2%)
Fuel- Air Mixture ---
------
---
Deposits
Oil Layers
Quenching
Crevices
(0.5%)
(1%)
(0.5%)
(5.2%)
rankcase (0.2%)
2.2%
4
1/3 Oxidized
4. %
5.1%
- Recycled -
1.7{%
e
I
Blow-by (0.6%)
- Recycled-
HC in Residual
Unburned
j
(1.1%) - Recycled -
1/3
Exhaust Oxidation
(0.8%)
e (0.1%)
In-Cylinder Oxidation I
2/3 Oxidized
1.5%
'-f
Exh. Valve
Leakag
2.1%
1/3
-
1.3%
:.
F
Fully Burned Exhaust
Engine- out HC (1.4%
Figure 1.2 Steady-state HC emissions flowchart [4]
25
di
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26
Spark plug
thread crevice
Lubricating oil
absorbtion and
desorblon
Valve seat
crevice
Cylinder head
gasket crevice
Piston ring
pack crevices
Bulk quenching within
flame front for lean
mixtures and high turbulence
Figure 1.3 Crevices in the combustion chamber [2]
2400
Piston Crevice Volumes
2000
-a-
E
0
0.613 cm3
-a- 1.665 cm3
-A- 1.675 cm3
1600
400-
E
LU800
400
0
0
2
6
4
8
Time (seconds)
Figure 1.4 Effect of crevice volume during warm-up [6]
27
10
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28
8000
S
I
6000
4000
8000
-
I
S
MID LOAD
LICHT LOAD
-
PFI
Ii
I
C
6000
~3000
K
CF,
2000
LA
0
0
20
100
80
60
40
TIME AFTER COtD START (SEC)
0
20
40
60
80
TIME AFTER COLD START (SEC)
b)
a)
Figure 1.5 Plot of total, engine-out HC emissions versus time after coldstart for two standard injection (PFI) and two pre-vaporized injection
(PV-CFI) starts. a) light load- 1200rpm, 0.74 bar BMIEP, fuel flow 0.9
g/sec. b) mid load- 1200rpm, 3.72 bar BMEP, fuel flow 1.75 g/sec [18]
160
140
E P o rt Wetting
El Cylinder Wetting
120 100 -
80 P
-
60 40
-
20
0
Warm
Warm-up
Figure 1.6 Liquid fuel distribution in the port and cylinder during coldstart and warm engine operation [17].
29
too
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30
Figure 1.7 Schematic of liquid fuel injector probe used by Stanglmaier
[14].
I-I
E-I
E-E
Figure 1.8 Liquid fuel injection location for direct injection of fuel used
by Stanglmaier [14][15].
31
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32
6000
5000
4000
E3000
2000
1000
0
E-E
I-E
E-1
I-I
Piston Top
Wetting Location
Figure 1.9 Effect of wall wetting location on engine-out HC's,
Stanglmaier [14].
33
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34
CHAPTER 2
EXPERIMENTAL METHOD
2.1 Nissan Test Engine
A Nissan SR20DE, 2L production four cylinder spark ignition engine was used for all
experiments. The 1991 Nissan Sentra engine has four valves per cylinder and a pent roof head
design with direct overhead camshafts. Previous to this study, the engine was modified to run on
only one cylinder with the other three cylinders motored.
The three motored cylinder intake
runners are separated from the intake plenum by a copper plate, the three runners are vented to the
atmosphere. The exhaust manifold was also modified to separate the firing cylinder from the three
motored cylinders.
This setup allows for precise control and measurement of the airflow and
exhaust gas of the firing cylinder. The complete engine technical specifications are listed in Table
2.1.
The Nissan engine is coupled to a 100HP Dynamatic dynamometer that is capable of
motoring the engine or absorbing the engine output when firing. Because of the intake manifold
modification, three of the four cylinders are motored at all times. Therefore, even while firing on
one cylinder, the engine load must increase to approximately 70% of maximum to overcome the
friction of the motored cylinders and produce positive power output. Regardless of load however,
the engine speed was maintained with a Digalog controller.
The fuel injector and port geometry of the engine used in these experiments is shown in
Figure 2.1.
The standard injector configuration consist of a 4-hole Nissan injector located
approximately 8.5 inches upstream of the intake valve. The injector has a 6 degree view of the
intake valve, the standard injector can only see 30%-40% of the total valve surface area.
Experiments were carried out to determine the emissions characteristics of the standard fuel
injection configuration.
2.2 Nissan Engine Control System
The Nissan engine was custom fitted with a 1999 Motec M4 engine control unit. This
programmable engine control unit (ECU) allowed for continuous monitoring of all critical engine
35
variables; manifold intake pressure (MAP), coolant temperature, intake air temperature, engine
speed, spark timing, fuel injection timing, and fuel injection pulse width. All measurement
devices used by the ECU were supplemented by a secondary and pre-existing measurement system
already on the Nissan engine. The Motec ECU allowed for
1
degree crank angle resolution for
spark timing and fuel injection timing across all operating speeds. Fuel pulse width resolution
could be controlled to within 0.1 ms (normal injector pulse width at 0.5 bar MAP is approximately
8 ms). The ECU allowed for the firing cylinder to be fueled with two injectors with different fuel
injection pulse widths. The programmable ECU also allowed for real time control of spark timing,
fuel injection amount, fuel injection timing.
2.3 Engine Measurement Systems
In order to enhance measurement precision of intake mass airflow (MAF) a 55 gallon
damping tank was placed ten feet downstream of the mass airflow sensor and six feet upstream of
the throttle. The intake airflow was measured using a Kurz, model 505-9A, mass air flow meter
and is displayed in units of grams per second. The Intake manifold pressure was measured at the
intake plenum entrance using a Data Instruments Model SA pressure sensor.
The fuel/air ratio was monitored using two independent universal exhaust gas oxygen
sensors;
a Horiba Model MEXA-11OX universal exhaust
gas oxygen sensor mounted
approximately 26 cm downstream of the exhaust port and a GM wide range lambda sensor
mounted 30 cm downstream of the exhaust port. The Horiba universal exhaust gas oxygen sensor
was used for all fuel metering while the secondary GM lambda sensor served as a check to ensure
that the Horiba sensor measurement values did not drift.
Engine coolant temperature was controlled through an external heat exchanger, cooled by a
chilled water supply. An external oil cooling circuit was also used to maintain the oil sump
temperature. The oil and coolant temperature were maintained at 800 C for all experiments with a
standard deviation of 1*C.
The engine was fitted with a 360 degree optical shaft encoder which outputs one digital
five volt pulse every degree of rotation as well as one digital five volt pulse per revolution. The
one pulse per revolution signal and the in-cylinder pressure data are used by the data acquisition
program to synchronize the data acquisition analysis program with the correct engine cycle
36
position. All data was sampled at a 1 degree interval using the output from a 360' degree per
revolution optical shaft encoder as the data acquisition trigger.
2.4 In-Cylinder Pressure Measurement
The in-cylinder pressure was measured using a Kistler 6051 piezoelectric pressure
transducer located in the cylinder head as show in Figures 2.2 and 2.3. The pressure transducer is
mounted in the cylinder head between the intake and exhaust valves of the number four cylinder.
The output signal from the pressure transducer was sent to a Kistler model 5010A dual mode
charge amplifier, the output signal of this amplifier was then sent to the data acquisition system.
In order to get an absolute pressure measurement from the pressure transducer, the pressure
must be referenced at one point in the cycle. The in-cylinder pressure was set to be equal to the
intake manifold pressure at BDCC. This is the recommended procedure used by Ford Motor
Company for low speed operation [23]. The in-cylinder pressure transducer was calibrated with a
dead weight tester over the pressure range of 1 to 35 bar to check for linearity. A typical pressure
calibration test is shown in Figure 2.4.
A further discussion of the in-cylinder pressure
measurement system used on this engine can be found in [24].
2.5 Hydrocarbon Measurement System
The Hydrocarbon emissions were measured using a Rosemount Model 402 Hydrocarbon
Analyzer. The Hydrocarbon Analysis is based on flame ionization.
The engine exhaust was
sampled from a ten gallon damping tank located 5 feet downstream of the exhaust port. The
sample was then drawn through a heated 15 foot line, the heated line temperature was set at 275'C
to prevent HC condensation in the sampling line. The Hydrocarbon Analyzer was calibrated using
nitrogen (0 ppm Cl) and propane calibration gas (4536 ppmCl).
Two different methods of
calibration were utilized: first the calibration gases were fed directly into the analyzer and second,
the same calibration gases were fed through the heated sampling line. This calibration technique
ensured that no residual hydrocarbons were in the sampling line. To reduce the probability of
residual Hydrocarbons in the sampling line the heated line was purged for five minutes with
nitrogen at the beginning and end of every experiment. At a fixed operating condition, the engineout hydrocarbon emissions day to day variation was found to be approximately 200 ppmC 1.
37
2.6 Labview Data Acquisition Program
The data acquisition system used in all experiments consisted of a Dell Pentium II
computer, National Instruments PCI-6025E multi-function 1/0 board, National Instruments BNC2090 BNC connector board, and a custom Labview 5.1 data acquisition program.
The
specifications for the Data Acquisition board are given in Table 2.2.
A Labview Data Acquisition program was written specifically for these experiments. This
program was designed with a continuously buffered data acquisition program, allowing the
program to sample and process a virtually unlimited number of engine cycles. The program
automatically calculated a continuous running average value for the Horiba Lambda sensor, the
GM Lambda Sensor, mass airflow (MAF), manifold pressure (MAP), and average in-cylinder
pressure based on the formula below.
New Average = [(Old Average * (Number of Samples - 1)) + New Value] / (Number of Samples)
This program also calculated the net indicated mean effective pressure (N-IMEP) for each
engine cycle. The in-cylinder pressure trace was displayed for each engine cycle and the N-IMEP
for each engine cycle was also plotted for comparison. The program monitored each engine cycle
for possible engine misfires, indicated by a negative N-IMEP. If an engine misfire was detected a
warning indicator turns red, and the number of misfires was counted.
After the desired number of engine cycles were acquired, the program output an average incylinder cycle pressure trace to a text file. The program also calculated the average N-IMEP, the
standard deviation of N-IMEP, peak cylinder pressure, and location of peak pressure.
The
complete list of program output values is listed in Table 2.3.
All input voltages to the data acquisition board were converted to the appropriate units prior
to being processed within the data acquisition program. The conversion formulas are listed in
Table 2.4.
2.7 Engine Operating Conditions
Liquid fuel flow into the engine cylinder is affected by many different engine variables,
therefore an experimental test matrix was developed. The review of previous work indicated that
the method of fuel injection and fuel targeting play a crucial role in the HC due to liquid fuel [3].
38
In standard port fuel injection, a liquid fuel film builds up in the intake port and around the valve
seat for both open and closed valve injection.
The baseline operating condition was chosen to be 1500 rpm with MAP equal to 0.5 bar,
this is close to the typical part load operating condition used by the auto industry. Engine speed
was not varied for this study because speed should only have a minor impact on liquid fuel
hydrocarbons.
The spark timing was optimized for maximum brake torque (MBT). MBT is
defined as the spark timing which results in maximum N-IMEP and corresponds to peak pressure
locations between 130 and 150 ATDC.
39
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40
Engine Type
Displacement / Cylinder (cm3 )
Clearance Volume (cm 3)
Bore x Stroke (cm)
Compression Ratio
Intake Valves
(34 mm Diameter / 10.2 mm Max Lift)
Exhaust Valves
(30 mm Diameter / 9.4 mm Max Lift)
Valve Overlap Period
4 valve/cylinder DOHC
Aluminum Head/Block
500
58.77
8.6 x 8.6
9.5
Open 130 BTDC
Close 235' ATDC
Open 4830 ATDC
Close 7230 ATDC
160
Table 2.1 Nissan engine specifications
I/O Board
Analog Inputs
Resolution
Sampling Rate
Input Range
Analog Ouputs
Analog Output Rate
Analog Output Range
Digital 1/O
Counter/Timers
National Instruments PCI-6025E
16 single ended/ 8 Differential
12 bits
200 kS/
+/- 0.05 to +/- 10 V
2
10 kS/s
+/- 10 V
32
2, 24-bit
Table 2.2 National Instruments data acquisition specifications
41
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42
Number of
cycles
acquired
Number of
Misfires
Average
N-tMEP
(Bar)
Standard
Deviation Of
N-tMEP
(Bar)
Maximum
Cylinder
Pressure
(Bar)
Location of
Maximum
Cylinder
Pressure
Average
Mass Airflow
(g/s)
Average
Manifold
Pressure
(Bar)
Average
Horiba
Lambda
Average
GM
Lambda 2
(deg. AT DC)
Table 2.3 Data output variables from data acquisition program
Signal
input
In-Cylinder Pressure
Top Dead Center Signal
Mass Airflow
Manifold Pressure
Horiba Lambda Sensor
GM Lambda Sensor
Slow RD
V - Volts
V - Volts
V - Volts
V - Volts
V - Volts
V - Volts
V - Volts
Conversion Formula
4.72*V - 0.55
1*V
100*V
0.36*V - 0.35
1*V - 0.55
-0.15
*V^2
-0.11*V + 1.1
-4020*V - 9.24
Output units
Bar
Volts
grams/second
Bar
Lambda
Lambda
ppm C1
Table 2.4 Conversion factors for raw voltage input to data acquisition
system
43
Average
HC
(ppm Cl)
(This page was intentionally left blank)
44
Figure 2.1 Standard intake runner and intake port geometry for the
Nissan Sentra engine
Figure 2.2 Top-view of in-cylinder pressure transducer location
45
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46
Kier
6051
3 mm
Piston #4
@ TDC
Recession = 6 mm
+--+
Figure 2.3 Side-view of pressure transducer location [24]
40
35-
y= 4.6828x - 0.3831
R2 = 0.9997
3025U)
20U)
15
105
U
0
2
6
4
Volts
Figure 2.4 Pressure transducer calibration data
47
8
10
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48
CHAPTER 3
EFFECTS OF ENGINE OPERATING PARAMETERS ON HC
3.1 Factors Affecting Liquid Fuel Distribution
Liquid fuel volume and distribution within the cylinder is dependent upon the intake
airflow and liquid fuel distribution within the intake port. The intake airflow is determined by:
intake port geometry, intake and exhaust valve timing, intake manifold pressure, and engine speed.
The liquid fuel distribution in the intake port is dependent upon engine coolant temperature, intake
valve temperature, injection timing, fuel injector targeting, injector spray angle, and fuel injector
droplet size and velocity.
3.2 Intake Airflow Processes
When the intake valve first begins to open, the cylinder pressure is higher than the intake
port pressure. This large pressure difference between the cylinder pressure and the intake manifold
pressure and the small intake valve gap as the valve begins to open, results in a sonic backflow of
exhaust gases into the intake port. This high-speed backflow of hot exhaust gases can redistribute
and vaporize the liquid fuel in the intake port, this effect is more significant at lower loads when
the pressure difference is higher. Experiments by Almkvist and Dahlberg found that this blowback
can blow fuel droplets far back into the intake ports at velocities up to 40 m/s [25]. An engine
simulation model developed at MIT estimates the distance which cylinder gas is blown back into
the intake port at 1200 rpm [26]. Figure 3.1 indicates that as the intake manifold pressure (MAP)
decreases the blowback gas volume and distance increase.
Previous experiments by Cheng,
estimated the mass flow rates through the intake valve at different intake pressures based on the
cylinder pressure data [27]. The three different intake processes can be seen in Figure 3.2.
When the blowback process has ended the piston motion induces a forward flow of air into
- the cylinder.
The intake process continues until after the piston reaches bottom dead center
(BDC). The compression stroke then begins to displace the charge mixture in the cylinder. At
higher speeds the inertia of the incoming airflow "rams" extra charge into the cylinder after BDC.
This allows for a higher volumetric efficiency at higher speeds but results in cylinder charge being
49
displaced as a backflow into the intake port at lower speeds. The intake valve typically closes 40
to 60 *after BDC resulting in a displacement blowback into the intake port at low operating speeds.
The intake and exhaust valve timing and resulting in-cylinder pressure data at the standard
operating point are plotted in Figures 3.3 and 3.4. By examining the in-cylinder pressure we can
estimate approximately where the three intake processes occur at the standard engine operating
condition of 1500 rpm and 0.5 bar MAP. The local minimum in the in-cylinder pressure trace at
point A, just after EVC is the estimated point where the blowback into the intake port stops and
the forward flow induction process begins. The induction process continues until a few degrees
after BDC. Approximately 50 after BDC (Point C) the cylinder pressure begins to rise during the
compression stroke and charge is displaced into the intake port.
As the manifold pressure increases, the initial overlap backflow of gases into the intake
port decreases. Increased engine speed reduces the displacement blow back at the end of the
intake process.
Qualitatively we can see from Figures 3.2 and 3.4 that the overlap and
displacement backflows are of approximately the same order of magnitude at the standard
operating conditions used in this study.
3.3 Effects of Spark Timing
To characterize the HC emissions sensitivity to various operating parameters for this test
engine, a test matrix was developed investigating the effects of spark timing, fuel air ratio
(lambda), fuel type, and mixture preparation.
The effects of spark timing on HC are shown in Figures 3.5 and 3.7, spark timing has a
modest impact on HC emissions for the Indolene fuel spark-timing experiments. The propanefueled experiments indicated a slight increase in HC with spark advance. With increasing spark
advance the maximum cylinder pressure increases and is located closer to TDC, causing more
unburned charge to be stored in the crevices. The spark timing also changes the in-cylinder gas
temperatures. The packing of unburned mixture may be of more significance to the propane fueled
experiments due to the fully mixed air/fuel charge.
A change in spark timing by 5 degrees
advanced or retarded causes a 13% increase or decrease in the average maximum cylinder pressure
and a shift of this peak pressure by 3.5 degrees for both fuel types. Figures 3.8 and 3.8 indicate
that the propane has a maximum N-IMEP when the peak pressure is located around 13 deg. ATDC
with a spark timing of 23 BTDC. Indolene fueled spark sweeps indicate a maximum N-IMEP
50
around 15 deg. ATDC corresponding to a spark timing of around 21 deg. BTDC. A second set of
propane and Indolene spark timing sweep experiments were controlled by matching the intake
manifold pressures for each fuel type. The mass airflow measurements of this experiment revealed
that the propane/air displacement causes a 3% decrease in volumetric efficiency and a
corresponding to a 5% decrease in N-IMEP. These spark timing experiments indicated that the NIMEP curve is relatively flat near MBT timing and that the HC do not change significantly for
spark timing, slightly advanced or retarded from MBT.
3.4 Effects of Relative Air/Fuel Ratio, Lambda
The effects of air/fuel ratio on HC emissions for propane and vaporized Indolene were
examined by running the engine at fuel rich and fuel lean conditions. The behavior of the HC
emissions are distinctly different on the rich and lean sides of stoichiometric engine operation as
shown in the HC and N-IMEP vs. Lambda curves shown in Figures 3.11 and 3.12. Under fuel-rich
engine conditions (Lambda <1), the HC level increases linearly as the charge mixture moves from
stoichiometric to rich engine operation. This is the expected trend as the burning mixture is airlimited, therefore an increase in the required fuel above stoichiometric results in a corresponding
increase in HC emissions. The N-IMEP curve shows a slight increase in N-IMEP as the charge
mixture becomes slightly fuel rich, with a peak around X=0.95. This peak in N-IMEP is due to
dissociation of CO 2 and H2 0 at the high temperatures following combustion, molecular oxygen is
present in the burned gases under stoichiometric conditions, so some addition fuel can be added
and partially burned [1].
On the lean side of stoichiometric the N-IMEP decreases linearly due to
the decreased amount of fuel available for combustion. The HC emissions decrease as the charge
becomes fuel-limited until a minimum is reached at X=1.1.
An increase in lambda above 1.1
causes an increase in HC as the total burn duration increases and the cyclic variability increases.
For a further discussion of the effects of relative air/fuel ratio refer to [1].
The trends in N-IMEP with lambda are identical for propane and pre-vaporized Indolene.
There is a slight difference in the slopes of the HC/lambda lines for the propane and the prevaporized Indolene. The difference between the two fuels in terms of HC is believed to be due the
different fuel properties of propane and Indolene. It is surprising however that there is not a
constant difference in the HC values.
This data suggest that the chemical composition and
51
corresponding differences in combustion between Indolene and propane should be considered in
the comparison of propane and Indolene experiments.
3.5 Coolant Temperature Experiments
Due to the increase in HC levels during cold engine operation, experiments were conducted
to characterize the engine under a prolonged engine cold-start/warm-up experiment.
These
experiments were conducted at the standard engine operating point and the engine coolant
temperature was slowly raised to standard engine coolant temperature over a 40 minute period.
Two test are shown for each engine fueling condition in Figure 3.14. The engine was fueled with
the standard Nissan injector with closed valve injection and with propane. The Nissan injector
fueled with liquid Indolene has a much higher HC level at cold operating conditions than at warm
engine conditions. The Nissan injector also has a change in the HC/coolant temperature slope
around 50'C. For coolant temperatures above 50'C, the propane and Indolene fueled experiments
have the same slope. The propane warm-up experiment demonstrates a linear dependence of HC
vs. coolant temperature. Because propane is not absorbed into the oil layer and is a gaseous fuel,
we cant postulate that the increase in HC with decreasing coolant temperature, for the propane
experiment, is primarily due to the increased quench layer thickness, changes in HC oxidation, and
increased crevices volumes. The higher HC's, at low coolant temperatures, for the Indolene fueled
experiment are believed to be due the increased amount of liquid fuel entering the combustion
chamber.
3.6 Propane and Indolene Fueling
To examine the emissions impact of running the engine on different stoichiometric
mixtures of propane and liquid Indolene a series of "liquid fuel sweep" experiments were done.
The engine was initially fueled by gaseous propane then a small amount of liquid Indolene was
injected and the propane flow was reduced to maintain a stoichiometric mixture. Figure 3.18
indicated that the HC level was directly proportional to the propane/Indolene mixture.
These
experiments were done with OVI timing using both the standard Nissan injector and the low-flow
BMW injector.
52
3.7 Effects of Fuel Injection Timing
By varying the fuel injection timing, the effects of OVI vs. CVI can be examined. The
effect of fuel injection timing on HC is shown in Figure 3.19, the data presented in Figure 3.19
consist of the average of two injection timing sweeps for each injector where each data point was
the average of 1000 engine cycles. The end of injection at the injector tip is plotted against the
actual valve timing. The HC peak before the IVO is indicative of the transport delay of fuel from
the injector tip to the intake. valve. There is a clear peak in the HC level for both the Nissan
injector and the low-flow BMW injector. This peak is the result of the maximum amount of liquid
fuel being transported into the cylinder, both the fuel droplets which are entrained into the intake
flow as well as liquid fuel film. Perhaps the more interesting observation that can be made from
this figure pertains to the closed valve portions of the graph. Figure 3.19 indicates that there is no
significant difference in HC's for "early CVI" (injection just after IVC) and "late CVI" (injection
just before IVO). Assuming that the liquid fuel distribution within the intake port is the same for
early and late CVI the HC's indicate that there is no significant difference in intake port fuel
evaporation for early vs. late CVI.
The large difference between open and closed valve injection is due to a variety of factors
such as the distribution of liquid fuel within the intake port, amount of evaporation in the port, fuel
droplet size, and blowback during IVO. These differences are likely to result in a difference in
mixture homogeneity within the cylinder and significantly different distributions of liquid fuel
within the cylinder.
53
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54
A
20
300
E
- 15
E
4)
a
0200 10
0
100
'
'
0
0
'0
0.2
0.4
0.6
0.8
1
Intake Pressure MAP (bar)
Figure 3.1 Computed intake backflow distance and volume, 1200 rpm,
cylinder displacement volume is 486cc[261
p= 0.85 atm
p =P0.6 atm
0.4 atm
10
CACA:
0W
am am. MM
W
W
urn 4M 40 "W
-50 0
@vo5
4W
VA
0r
50
MW
-O MA
-O WM
10
MOP
0
OW mm
0
5
2W0
250
BC
10E-3
1050
0
50
too
150
CA Dogr.
Figure 3.2 Calculated mass flow rates through intake valve at 0.4, 0.6,
and 0.85 Bar MAP, [27]
55
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56
5
4.5
4
3.5
3
(D
L(D
EVO
EVO
2.5
2
1.5
1
EVC
0.5
1VC
0
0.00
0.20
0.40
1.00
0.80
0.60
Cylinder VolumeNmax
Figure 3.3 Intake and exhaust stroke in-cylinder pressure data
2
1.8
s\
1.6
EVO
IVO
I-
.0
a)
1~
1.41.2
U,
U,
a)
aL.
a)
V
0
1-
Cylinder Pressure Setpoint
0.8 0.6
0.4
AB
0.2
C
0
100
150
200
250
300
350
400
450
500
550
600
Crank Angle (deg ATDC)
Figure 3.4 Cylinder pressure during intake and exhaust. Point Aestimated end of blow back, Point B- forward induction, and Point Cend of induction-displacement backflow
57
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58
4000
3500
-
3000
-
-*- Indolene
-+- 50/50
2500
-
---
Propane
EL 2000 CL)
:
1500
-
1000
-
500
-
Retard
Advance
-p
4-
0-8
-6
-4
-2
0
2
4
6
8
Spark Timing From MBT
Figure 3.5 Spark timing vs. HC with matched N-IMEP for propane and
Indolene
24
23
~22
-o- Indolene
-+50/50
= 21-
-A- Prop ane
~20
S19
E
.
18
-
W"',
17-
Retard
16-
4
-
Advance
f1 -
15
-8
-6
-4
-2
0
2
4
6
Spark Timing From MBT
Figure 3.6 Maximum cylinder pressure vs. spark timing with matching
of N-IMEP for propane and Indolene Mixtures
59
8
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60
4000
-
* Indolene
3500 -
A
3000 -
Propane
2500 E
a. 2000 CL
=
1500 1000 -
500
A
A
A A AAAAAAA
A
A
Advance
-
30o
015
25
20
35
30
Spark Timing (deg BTDC)
Figure 3.7 Spark timing vs. HC with Nissan injector and propane
3.2
e Indolene
A
Propane
3.1 -
CL
wL 3....-......
A
A
2.9-
A30
2.8
15
20
25
30
35
Spark Timing (deg BTDC)
Figure 3.8 Nissan Injector and Propane fueled N-IMEP vs. spark timing,
with MAP = 0.5 bar
61
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62
3.2
e
Indolene
A Propane
3.1
3-
z
2.9-
2.8 -0
15
10
5
20
Location of Peak Cylinder Pressure (deg ATDC)
Figure 3.9 Nissan Injector and Propane fueled N-IMEP vs. location of
peak cylinder pressure, with MAP = 0.5 bar
20
A Indolene
A
0
I-
e Propane
0)
C)
-o 154C)
1..
C,,
C)
e
a-L..
0 eA
eA
0
A
C)
a0
10 +
C
.2
*6-'
C.)
0
-J
5
15
20
25
30
Spark Timing (deg BTDC)
Figure 3.10 Location of peak cylinder pressure vs. spark timing
63
35
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64
4000
Pre-Vaporized Indolene
AA
3500 -
* Propane
A
A A
A
3000 -
2500
E
CL
0.
2000
*A
1500
1000
500
0
0.75
0.8
0.85
0.9
0.95
1
1.05
1.1
1.15
1.2
1.25
Lambda
Figure 3.11 HC vs. Lambda for propane and pre-vaporized Indolene
experiments
3.1
A Pre-Vaporized Indolene
e Propane
A
3
4
2.9
4
4
S
w
2.8 f
2.7
4L
2.6
0.75
0.8
0.85
0.9
0.95
1
1.05
1.1
1.15
1.2
Lambda
Figure 3.12 N-IMEP vs. Lambda for propane and pre-vaporized
Indolene
65
1.25
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66
15
*
Pre-Vaporized Indolene
A Propane
104
0)
o'
s
*
Ae*
A
0
5
0
0.75
0.8
0.85
0.95
0.9
1
1.1
1.05
1.15
1.2
1.25
Lambda
Figure 3.13 Effects of Lambda on cyclic variability
6000
55001
'50001
4500
4000
-
3500
-
0. 3000
-
E
0.
2500
-
2000
-
1500
-
I..
A
Op
A
A
AA A
A AhAA A AA
AAA
A~
Propane
1000500
Closed Valve Nissan Injector
-
00
20
40
60
80
100
Coolant (deg C)
Figure 3.14 HC engine warm-up experiments for propane and standard
Nissan injection of Indolene
67
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68
3000
2800 2600 2400
ANissan Injector
*BMW Injector "Low-flow"
-
A
2200 -
E
0.
CL
A
2000 -
I
A
A
1800 1600
-
00
1400 1200 1000
0
100
80
60
40
20
Percent of Required Indolene
Figure 3.18 Liquid fuel sweeps with Nissan and BMW injectors, OVI
4000
-i-- Nissan Injector
--o-- BMW Injector
3500
Transport Delay
3000
2500
E 2000
GL
C)
1500
IVO
BMW P.W.
1000
Niss. P.W.
500
.
0
0
I I
I . I
100
.
I I .
200
I
1
300
.1
1
400
1
1 i
500
I
.
.
I .!
I
I
I .
600
End of Fuel Injection (BTDCC)
Figure 3.19 Injection timing vs. HC for standard Nissan Injector and
the low-flow BMW injector. Injection timing corrected for 175 degree
transport delay.
69
i
700
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70
CHAPTER 4
LIQUID FUEL INJECTOR PROBE EXPERIMENTS
4.1 Overview of Liquid Fuel Injector Probe Experiments
To simulate the liquid fuel flow from the valve seat region into the cylinder, a specially
designed injector probe was constructed to inject controlled amounts of liquid fuel onto the port
wall close to the valve seat. By fueling the engine on a gaseous fuel, either propane or prevaporized Indolene, and injecting liquid fuel close to the valve seat, we can examine the effects of
liquid fuel entering the cylinder at different locations around the valve seat. These experiments
were carried out with both OVI and CVI to assess the differences that residual blowback and
evaporation produce.
4.2 Pre-Vaporizing Injector
The effects of operating the engine on two different fuel types (propane and
Indolene) simultaneously at differing mixture ratios makes the interpretation of HC data difficult at
best. The two most notable problems of operating on two fuel types are the different deposit
forming characteristics of the fuels and the negligible absorption of propane into the oil layer. The
heavier components of the Indolene fuel have a tendency to form deposits on both the intake
valves and in the combustion chamber. Propane, a single component gaseous fuel, did not form
any significant deposits on the intake valve and only minor deposits in the combustion chamber.
A more significant difference between the fuels is the absorption of fuel into the oil layer.
Experiments by Gatellier and Herrier verified that there is no significant absorption of propane
into the oil layer [30]. However, Indolene will be absorbed into the oil layer, causing the oil layer
to act as a storage medium for unburned fuel. It was found that after running the engine on
Indolene for an extended period of time the HC for the next set of propane experiments exhibited a
higher than normal level of HC. This higher level of HC's was believed to be due to the fuel
stored in the sump oil and desorbing from the oil layer on the cylinder liner. To circumvent these
problems with fuel deposits and oil layer effects, a pre-vaporizing gasoline injector was developed
and used.
71
The pre-vaporizing injector consist of a low-flow injector mounted at one end of a heated
brass tube. An air supply line injects air into the brass tube, inducing a swirling flow through the
tube. The liquid fuel injected onto the hot brass surface is evaporated and transported out of the
injector via the swirling airflow. This pre-vaporizing injector was based on the design of an earlier
pre-vaporizing injector developed at the National Engineering Laboratory in England [31]. The
schematic of this injector is shown in Figure 4.1.
The swirling airflow was supplied by a pressurized oil free air supply with the air-assist
flow rate controlled by a critical flow orifice. The orifice used was an O'Keefe Controls Metal
Orifice Number 38.
The mass airflow rate was controlled by varying the pressure difference
across the critical flow orifice. This mass airflow was checked by running the engine at a fixed
manifold pressure (0.5 bar) and varying the mass airflow of the air-assist. The total engine airflow
was measured by two different methods: 1) the total airflow was estimated using the fuel mass
flow and the lambda measurement, 2) total airflow was estimated by adding the mass airflow
measured at the air intake and adding the air-assist mass airflow based the orifice calibration
Tables.
This experiment to characterize the mass airflow of the air-assist at different supply
pressures is shown in Figure 4.2. At the standard air-assist operating point with a supply pressure
of 10 psi there is only a 4% difference between the two mass airflow measurement systems for the
total airflow. Also note that the HC level for different air-assist supply pressures does not change
significantly over the test range. The supply pressure for the air-assist was set at 10 psi (14.5% of
total mass airflow) for all pre-vaporized experiments.
4.3 Liquid Fuel Injector Probe
The liquid fuel injector probe consist of a low-flow injector connected to a 1/16"
inner diameter Teflon tube. Each of the fuel probe tubes were approximately 14" in length and are
epoxied to the intake runner wall. The end of the fuel probe tubes were mounted
" behind the
intake valve seating surface and were aligned such that they were perpendicular to the valve seat.
Three injector probe tubes were permanently mounted in the intake port in the locations shown in
Figure 4.3.
To ensure precise injection timing could be achieved with the liquid fuel injection probes, a
series of flow visualization studies were conducted. The visualization experiments consisted of a
free hanging 14" Teflon tube injecting liquid fuel into a beaker. The injector was pulsed at a rate
72
of 750 pulses/minute and a fuel pulse width of 2.8 ms, this corresponds to 32% of the total fuel
required to fuel the engine at the standard operating point of 1500 rpm, 0.5 bar MAP. A strobe
light allowed visualization of the probe flow at different locations in the engine cycle.
This
visualization study indicated that there is a transport delay of 185 degrees from the start of
injection until the liquid fuel begins to exit the fuel probe tip and a 260 degree delay from the end
of injection until the liquid fuel stops flowing from the fuel probe tip.
Therefore, a signaled
injection pulse width of 25 crank angles results in an effective pulse width of 100 crank angles at
the fuel probe tip. The results of this visualization study are shown in Figure 4.4.
4.4 Liquid Fuel Probe Injection Timing Sweeps
A fuel injection timing sweep was done at each probe location. The engine was fueled
with 36% of the required Indolene injected through the fuel probe and the balance of the required
fuel was supplied by propane. The actual end of injection is adjusted to match the actual engine
cycle position based on the visualization study and the phasing adjustment shown in Figure 4.4.
The effects of injection timing on HC's and N-IMEP at the probe "B" location are shown in Figure
4.5. This graph indicates a clear step type increase in HC emissions during open valve injection.
The rise and fall time of this step increase is approximately equal to the injection duration. As the
fuel injector begins to inject fuel during IVO, the HC's begin to increase. The HC's continue to
increase until all the liquid fuel is being injected during open valve injection. The HC's begin
decreasing at BDC when the intake airflow ceases.
The increase in HC's and corresponding
decrease in burned fuel results in a decrease in N-IMEP during open valve injection. This increase
in HC and decrease in N-IMEP during OVI was found at all three probe locations as can be seen in
Figures 4.5, 4.7, & 4.9. It should also be noted that there is a significant increase in the cyclic
variation of N-IMEP during open valve injection. This is shown by the variation in covariance
defined as:
Covariance = Standard Deviation of N-IMEP *100
Mean N-IMEP
Figures 4.6, 8, &10 indicate that the covariance follows closely with the HC emissions
trend for all three probe locations and injection timing. The comparison of the HC's for all three
probe locations indicate that the average open valve and average closed valve HC's are different
for each probe location. This set of injection timing sweep experiments had a large amount of
73
experimental scatter during the open valve portion of the timing sweep. There is however, a clear
difference in the closed valve injection HC's between the probe locations.
Probe "A"
demonstrated the highest HC level during closed valve injection, with probe "C" having the lowest
HC's. In comparing the covariance of N-IMEP for the different probe locations we find that the
covariance is higher during open valve injection for all three probe locations, as can be seen in
Figure 4.12. In examining Figure 4.12 we see that the covariance values are similar for all three
probe locations. This injection timing sweep and the probe visualization study verify the accuracy
of the liquid fuel injector probe timing.
4.5 Fuel Probe Indolene/Propane Fuel Sweeps
The impact of the injected liquid fuel mass was studied by injecting different amounts of
liquid fuel to examine the corresponding effect on the HC's at each probe location. This was done
for both OVI and CVI and fueling the engine on a combination of propane and Indolene. The
result of this fuel sweep for both open and closed valve injection on HC's is shown in Figures 4.13
and 4.14.
The closed valve injection of liquid fuel at each probe location shows a linear
relationship between HC's and the percent of injected Indolene. This linear relationship was also
found when doing a similar fuel sweep with the standard injector (see Figure 3.18). Once again,
there is a clear difference in the HC between each probe location with probe "A" having the
highest HC's and probe "C" the lowest. Also note that the slope of the HC vs. percent Indolene is
different for each probe location. A problem arose when this experiment was carried out for OVI,
the engine began misfiring for probe fueling above 22%. At this propane setting we are within the
lean misfire limit even if no liquid fuel is being injected. The reasons for the engine misfires at
this operating condition remain unclear, and prompted the use of a 100% Indolene fueling with a
pre-vaporized injector.
4.6 Fuel Probe Fuel Sweeps
The utilization of the pre-vaporized injector in conjunction with the liquid fuel probe
closely resembles the fueling in a standard PFI system. By using the same fuel type for both the
pre-vaporizing injector and the liquid fuel probe, the interpretation of the experimental data was
simplified.
The effects of effects of various amounts of fuel being delivered at different
74
circumferential locations around the valve seat were studied by fueling the engine on different
proportions of liquid fuel injected at the valve seat at each probe location.
The cyclic variability or covariance of these experiments was much lower than the initial
propane/Indolene experiments.
However, the cyclic variability increased as the proportion of
liquid fuel increased and was larger for OVI than for CVI, see Figure 4.15. It should also be noted
that there is no significant difference in the cyclic variability between the different probe locations.
This indicates that the difference in HC's between the probe locations is due to factors other than
differences in cyclic variability.
These experiments indicate that for all amounts of liquid fuel injected at the valve seat, the
fuel delivered closest to the exhaust valve (probe "A") results in the highest engine-out HC's,
while the location farthest from the exhaust valve (probe "C") has the lowest HC's. Figures 4.17
and 4.19 show the effect of liquid fuel injected at the valve seat during CVI. Figures 4.18 and 4.20
show the effect of liquid fuel injected at the valve seat during OVI.
A method of interpretation of the experimental probe data was developed based on the
mass of the HC's. The following relationship was assumed:
MHC/Muel = XHC =
y fliq +(1 -y)yvap
eq. 1
MHC - mass of engine-out HC's
Mfuel
- mass of total injected fuel
XHC
- mass fraction of fuel emitted as HC's
y
- mass fraction of fuel injected into port as liquid
Wfiq
- mass fraction of liquid fuel in the cylinder emitted as HC's
Nfvap - mass fraction of vaporized fuel emitted as HC's
Also realizing that a certain amount of the liquid fuel entering the cylinder is vaporized prior to
combustion, (z - mass fraction of fuel probe "liquid" not vaporized in the cylinder prior to
combustion) we can expand equation 1.
MHC/Muel = XHC =
YZ
Whiq
+(1-yZ)Wvap
75
eq. 2
The values of 4fiiq and V4vap were estimated by a linear regression analysis of the experimental data
based on equation 2 and assumed values for z.
Previous work by Meyer [3], suggest that
approximately 50% (z = 0.5) of the liquid fuel entering the cylinder is vaporized prior to
combustion. The results of this linear regression analysis are given in Table 4.1 for z = 1, 0.5, and
0.1. The mass fraction of injected vaporized fuel emitted as HC's is approximately constant for all
probe locations and all z values.
By examining the ratio of
iq/Wvap
we find that during OVI the liquid fuel fractions of probe
locations A, B, and C compared to the pre-vaporized liquid fuel fractions are 7 times, 5 times, and
3 times the value of the pre-vaporized liquid fuel fraction. During CVI, the liquid fuel fraction of
probe locations A, B, and C are 5 times, 2 times, and 1.3 times the pre-vaporized fuel fraction.
This analysis indicates that a given amount of liquid fuel entering the cylinder as liquid at probe
location "A" (closest to the exhaust valve) will result in 7 times the engine out HC's of an equal
amount of fuel entering the cylinder as vaporized fuel. While a given amount of liquid fuel
entering the cylinder as a liquid at probe location "C" (farthest from the exhaust valve) will result
in 3 times the HC's of an equal amount of vaporized fuel
Table 4.1 also shows the same trend as Figures 17, 18, 19, & 20; probe "A" has the highest
mass fraction of fuel emitted as HC's for OVI and CVI, while probe "C" has the lowest mass
fraction of fuel emitted as HC's. For comparison purposes the mass fraction of fuel emitted as
hydrocarbons for the standard Nissan injector with CVI and OVO and the mass fraction of
hydrocarbons with pre-vaporized Indolene fueling are given in Table 4.2.
Figures 4.21 and 4.22 compare the N-IMEP vs. percent of fuel injected as liquid. Figure
4.21 indicates that there is no significant change in N-IMEP for CVI with different amounts of
injected liquid fuel. Figure 4.22 shows a consistent decrease in N-IMEP with increased amounts
of liquid fuel for OVI. The N-IMEP reaches a minimum around 33% of the fuel being injected
through the fuel probe, the N-IMEP decreases by 3.3%. This decrease in N-IMEP is believed to be
due primarily to the increasing cyclic variability shown in Figure 4.23.
4.7 Comparison of OVI and CVI with Injector Probes
The injection timing sweep experiments with the fuel injector probes indicate that there is
a distinct step type increase in the HC's during OVI. This also corresponded well with an increase
in the cyclic variability or covariance. The differences in HC between OVI and CVI are similar for
76
all the probe locations, the initial interpretation was that this indicated a similar rate of evaporation
of fuel at all probe locations for OVI vs. CVI. However, on further examination we find that by
comparing "early CVI" just after IVC to "late CVI" there is no significant difference in HC and we
can assume no significant difference in evaporation.
Stanglmaier [14] also found that with his
direct injection fuel probe experiments the HC's were only weakly dependent on injection timing,
stating that there is not enough time in the engine cycle for any significant vaporization to occur.
The difference between OVI and CVI HC's is believed to be due to blowback from the
cylinder just after IVO. This blowback of hot residual gases redistributes the liquid fuel in the
intake port and aids in the fuel air mixing process for CVI. The blowback flow velocities around
the valve seat will all be approximately equal and therefore have the same effect at all three probe
locations. With OVI all of the liquid fuel injected at the valve seat enters the cylinder from one
location and leads to a higher (compared to the CVI case) variation the in the fuel/air mixture.
This higher variation in the mixture uniformity results in a higher cyclic variability for the OVI
compared to CVI. Figure 4.16 compares the HC values for OVI and CVI at probe locations "A"
and "C". The difference between the OVI and CVI HC's appears to be the same for all probe
locations and indicates that blowback has a similar effect at all three probe locations.
4.8 Experimental Fuel Probe Comparison with PFI
A simplified comparison of the experimental probe data can be made with standard port
fuel injection using equation 2. Experimental data for the standard Nissan injector with closed
valve injection indicated a total engine-out HC fuel fraction of
XHC
= 0.0157±0.0007.
For
simplification, the standard Nissan injection is approximated by an average of all three injector
probes with OVI. Assuming 50% of the liquid fuel entering the cylinder is vaporized prior to
combustion [3]. Using equation 2, an estimate of the amount of liquid fuel entering the cylinder
with standard closed valve Nissan injection can be made using the liquid fuel probe data.
MHC/Mfuel = XHC =
YZ
liiq +(1 -yZ)vap
XHC
= 0.0157 experimental data for standard Nissan injector
Wgiq
=(WA+VB+WC
Vvap
= 0.0144 see table 4.1 - vapor HC fuel fraction
)/3 = 0.0671 see table 4.1 - liquid HC fuel fraction (z = 0.5)
77
z
= 0.5
0.0157
= y(O.5)(0.0144) + (1-y(0.5))(0.0671)
based on [3]
This analysis indicates that approximately 5% of the total injected fuel mass enters the
cylinder as liquid with standard Nissan closed valve fuel injection. Noting that
(0.05)(0.5)(0.0671)
XHC, liquid = YZfliq=
= 0.0017, the mass fraction of total injected fuel exiting the engine as
hydrocarbons due to liquid fuel entering the cylinder is estimated. By using the oxidation values
in the HC emissions flowchart in Figure 1.2 an estimate of the total injected fuel escaping
combustion due to liquid fuel can be obtained. The percent of fuel escaping combustion do to
liquid fuel is approximately XHC,liquid(3 12 )( 312 )( 31 2 ) *100%= 0.6%. Previous research indicates
that approximately 1.2% of the fuel injected escapes combustion due to liquid fuel [4].
This
simplified comparison of the liquid fuel probe experimental data to standard port fuel injection
indicates that 0.8-0.2% of the total fuel escapes combustion due to liquid fuel, based on this
analysis.
These values are only approximate but they indicate that the liquid fuel probe
experiments are a realistic approximation of liquid fuel transport into the cylinder of a standard
port fueled SI engine.
4.9 Summary
The liquid fuel injector probe in conjunction with the pre-vaporized injector is an accurate
method for the simulation of the standard PFI system. Strobe visualization of the liquid fuel probe
injection and fuel injection timing sweeps have validated the accuracy of the injection timing and
control of the liquid fuel injector probe.
Both the propane/Indolene and the Indolene-only
experiments indicated that there is a distinct difference in the HC behavior at the different fuel
probe locations. Location "A" is the highest HC for both OVI and CVI and probe "C" is the
lowest.
The difference in the HC's between OVI and CVI is similar for all probe locations
suggesting that the blowback effect is roughly the same at all locations around the valve. The
cyclic variability is also similar for each probe location but higher for OVI than for CVI.
78
Z=1
Fraction
Fuel
HC
Liquid
Z=O.5
Liquid HO.Fel.Facton
Vapor HC Fuel Fraction
Fuel Fraction
Vapor HO Z=0.1
Liquid HC Fuel Fraction
B-CV
0.0240
A-CV
0.0432
A-CV
-B-CV
C-CV
0.0171
A-OV
0.0561
B-OV
0.0392
C-OV
0.0270
C-CV
A-OV
B-OV
C-OV
.
0.
0.0142i
0.0145.
0.0193
0.0149
0.0975
0.0146
0.0142
B-CV
0.012
0.0149.
C-CV
0.019
0.0146.
A-OV
0.049
0.0145'
A-CV
0.014
0.0648
0.0137
0.0391
0.0149
0.013T7
B-OV
0.064
0.0149
C-OV
0.0139
Table 4.1 HC fuel fraction values based on linear regression analysis of
experimental data with different assumed Z values. Z- is defined as the
fraction of injected liquid fuel enter cylinder as liquid (Z=O all injected
liquid fuel is vaporized prior to combustion).
+/- 95% Confidence Interval
Average
Nissan Injector CVI
0.0157
0.0007
Nissan Injector OVI
0.0211
0.001
Pre-Vaporized Indolene
0.0149
0.0007
Table 4.2 HC fuel fraction values for standard engine operating
conditions
79
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80
Low-flow injector
10" heated section
LP
Air-Assist
Heated Brass Tube
Figure 4.1 Schematic of pre-vaporizing injector
3500
50
A
U)
40 -
Based on Fuel Flow and Lambda
* Based on MAF and Orifice Mass Flow Tables
* HC
3000
2500
0
30 +
0
20
-
2000
E
A
Standard Operating Condition
10 psi supply pressure
14.5% of Total Airflow
A
0.
0
1500
A
A
0
1000
u
-0
0
10
nA
-
500
114-
Supply line open to Atmosphere
0
0
0
20
10
30
40
Delta P across Air Assist Orifice
Figure 4.2 Air-Assist mass airflow estimation, intake manifold pressure
set at 0.5 bar
81
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82
Liquid Fuel Injection locations
D
H
B
4
AJ
Intake
Exhaust
Airflow
Intake
Exhaust
Cross Section D
Top View of Cylinder
Figure 4.3 Liquid fuel injector probe locations
100
flow rate at injector tip
75
0
L
flowrate at probe tip
50
260 degree delay of end of Inj.
25
185 degree delay of start of lnj.
0
0
25
50
75
100
125
150
175
200
225
250
275
300
325
Relative Crank Angle
Figure 4.4 Liquid fuel injection delay through fuel probes, based on
liquid fuel probe visualization study
83
350
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84
3.4
4000
3500
-A-- Standard Nissan Injector
~~Probe B - HC
-
- - Probe B - N-IMEP
-IVO
3000
-
2500
-
3.2
3
cc
2.8
z
E
2000 -
Ivo
C)
~ p.
-
1500 +
p.
11-,
p.-
1000
P robe
nj. Duration
2.6
500
0
0
100
200
400
300
700
600
500
End of Injection Timing (after TDC compression)
Figure 4.5 Probe B Liquid fuel probe injection timing sweep, effect of
fuel injection timing on HC and N-IMEP, 36% Indolene, balance
Propane
20
4000
---- Probe B - HC
3500 -
- - - - Covariance of N-IMEP
IVO
15
3000
-e
2500
E
C-
CL
I
'
.4
p.
~.
0
2000
s
C)
p.
IV
*f*
hh4
1500
V
ep.
10
'
4'*
'4,
0
-5
1 000 f
500
1L
0 1.
0
100
200
400
300
500
600
40
7 00
End of Injection Timing (after TDC compression)
Figure 4.6 Probe B Liquid fuel probe injection timing sweep, effect of
fuel injection timing on HC and Cyclic Variation of N-IMEP
(Covariance) , 36% Indolene, balance Propane
85
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86
4000
3.4
-r
1
-A- Probe A - HC
- - - -
3500
3000
Probe A -N-MEP
IvO
3.2
-
2500 -
3
Z
'U
E
C. 2000
Ivo
C.
2.8
-*-
1500
.'
1000
U
-U
Probe
ij. Duration
2.6
500
0
100
0
400
300
200
500
!0
72.4
600
700
End of Injection Timing (after TDC compression)
Figure 4.7 Probe A Liquid fuel probe injection timing sweep, effect of
fuel injection timing on HC and N-IMEP, 36% Indolene, balance
Propane
1
I
4000
20
IvO
Probe A - HC
Covariance of N-IMEP
-A----
3500
1
15
3000 1
-LU
2500
C-
-
N" O
0
E
0
E 2000 0
I
= 1500 - - -
-. 1
*~
- 5
,'ge
*
00
,,,
Probe
Inj. Duration
1000
00
14
0
0
100
200
400
300
500
600
7
End of Injection Timing (after TDC compression)
Figure 4.8 Probe A Liquid fuel probe injection timing sweep, effect of
fuel injection timing on HC and Cyclic Variation of N-IMEP
(Covariance) , 36% Indolene, balance Propane
87
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88
4000
3.4
--- - - -
3500
Probe C - HC
Probe C -N-IMEP
IvO
-3.2
3000
2500
E
3
2000
Ivo
C)
-
1500
-
S2.8
- - -.
~---
1000'[
Probe
ij. Duration
-2.6
500-2.4
0
0
200
100
400
300
500
600
700
End of Injection Timing (after TDC compression)
Figure 4.9 Probe C Liquid fuel probe injection timing sweep, effect of
fuel injection timing on HC and N-IMEP, 36% Indolene, balance
Propane
20
4000
--3500
Probe C - HC
- - - - Covariance of N-IMEP
-
IvO
15 ^
3000
I.
(U
2500
E 2000
a.
C)
U
,
10
r\/O
t
z
0
1500
50
1000
0
0
100
200
400
300
500
600
700
End of Injection Timing (after TDC compression)
Figure 4.10 Probe C Liquid fuel probe injection timing sweep, effect of
fuel injection timing on HC and Cyclic Variation of N-IMEP
(Covariance) , 36% Indolene, balance Propane
89
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90
4000
A
3500 +
3000
-
2500
-
A-A- A-
A
A
A:
eA-AA
E 2000
1500
NO
-
1000 -
-A -- Probe A
-Probe B
500
-C-- Probe C
-
Ivo
I-
0
400
300
200
100
500
600
700
End of Injection Timing (after TDC compression)
Figure 4.11 Comparison of HC's for probes A,B, and C. 36% Indolene,
balance Propane
20
-- A- Probe A
-4-Probe
B
-L- Probe C
15
A
IL
wU
z
'4-
0
10
---V
A~ .A
0
-'V
0
5
0
0
100
200
400
300
500
600
End of Injection Timing (after TDC compression)
Figure 4.12 Comparison of Covariance for probes A,B, and C. 36%
Indolene, balance Propane
91
700
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92
4500
A Probe A
4000
e Probe B
n Probe C
3500
A
3000
2500
E
.E 2000
A
1500
A
1000
500
0I
10
5
15
20
25
45
40
35
30
50
Percent Indolene
Figure 4.13 Indolene/propane fuel sweep, CVI. end of injection 70
deg.after TDCC, 280 deg.before IVO
4500
Probe A
.Probe B
A
A
4000
A
A
A
A
I Probe C
3500
3000
E
CL)
A
2500
1 or more misfires/ 1000 cycles
2000
1500
1000
500
2-4 misfires/ 1000 cycles
0
0
5
10
15
30
25
20
Percent Indolene
35
40
45
Figure 4.14 Indolene/propane fuel sweep OVI, End of Injection at 410
deg. after TDCC, 60 deg. after IVO
93
50
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94
0.45 -A- Probe A
Probe B
0.4
-
0.35 -
Open Valve
-u-Probe C
0.3 -
0.25 -
'
CL
0.2 -
cc Z
S0 0.15
Closed Valve
0.1 0.05 0
0
5
10
15
25
20
30
35
40
45
Percent Indolene from probe, balance is Vaporized
Figure 4.15 Comparison of standard deviations of N-IMEP for liquid
fuel probe sweep
5000
A Probe A - CV
4500
* Probe C - CV
A Probe A - OV
4000
o Probe C - OV
3500
E
CL
-
A
3000
0
A
2500
2000
0
5
10
15
20
1
25
30
35
40
Percent of Liquid Indolene from Probe, balance is Vaporized
Figure 4.16 Comparison of OVI and CVI HC's for probes A and C
95
45
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96
5000
A
4500
Probe A
* Probe B
4000
m Probe C
3500
3000
UA
E 2500
0.
U
2000
1500
1000
500
0
0
5
10
15
25
20
30
35
40
45
Percent of Liquid Indolene from Probe, balance is Vaporized
Figure 4.17 Liquid fuel probe sweep CVI HC's, end of injection 70
deg. after TDCC, 280 deg. before IVO
5000
A
4500 -
Probe A
e Probe B
A
- 4000
m Probe CA
3500
3000 E 2500
20
3 2000
1500
1000
500
0
0
5
10
15
25
20
30
35
40
45
Percent of Liquid Indolene from Probe, balance is Vaporized
Figure 4.18 Liquid fuel probe sweep OVI HC's, End of Injection at 410
deg after TDCC, 60 deg. after IVO
97
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98
0.00025
-MA
(D
0.0002
N
Probe A
A
U)
Probe B
-
0
Probe C
CLe
0.00015
-L
y =2.75E-02x
A
0)
0.0001
0
0
y
0.00005
7.77E-03x
.
0
0.001
0.004
0.003
0.002
y = 2.33E-03x
0.005
0.006
Mass of Fuel Injected through probe (g/cycle)
Figure 4.19 Effect of liquid fuel on exhaust HC's, CVI
0.00025A
(n
Probe A
A
Cu
0.0002--
0 Probe B
y
=
4.05E-02x
* Probe C
0
C.
A
0.00015
0
0
mA
y = 2.19E-02x
0.0001
y =1.22E-02x
0g
0.00005
00
0
0
0.001
0.003
0.002
0.004
0.005
Mass of Fuel Injected through probe (g/cycle)
Figure 4.20 Effect of liquid fuel on exhaust HC's, CVI
99
0.006
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100
3.1
A
Probe A
* Probe B
0
3.05
A*0
AA
A
A
£
N
I-
0
U
3
Cu
*Probe
U
U
C
.
M
0~
w
2.95
z
2.9
2.85
15
20
25
30
40
35
45
Percent of Liquid Indolene from Probe, balance is Vaporized
Figure 4.21 Effect of liquid fuel probe mass on N-IMEP, CVI
3.1
A Probe A
5
3.05-
Probe B
* Probe C
A
3.
A
A
wU
E
2.95
-
AA
2.9
2.85
15
20
25
30
35
40
Percent of Liquid Indolene from Probe, balance is Vaporized
Figure 4.22 Effect of liquid fuel probe mass on N-IMEP, OVI
101
45
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102
3.2
3.15 -
A
3.1 AAA
3.05
a(CU
wn
A
3
2.95
y = -0.984x + 3.2639
A
2.9
2.85
2.8
0
0.05
0.1
0.15
0.2
0.25
0.3
0.35
0.4
0.45
Standard Deviation of N-IMEP (bar)
Figure 4.23 Correlation of N-IMEP with the standard deviation of NIMEP for all liquid fuel probe sweep data
103
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104
CHAPTER 5
SUMMARY AND CONCLUSIONS
5.1 Liquid Fuel Transport into Cylinder
Liquid fuel which is injected into the port is transported into the cylinder in three possible
forms; fuel vapor, fuel droplets, or as a fuel film on the port wall. Various researchers have
examined this fuel transport from the port to the cylinder [3]. During closed valve injection (CVI)
most of the liquid fuel is in the form of a fuel film on the port walls and intake valve prior to IVO.
Open valve injection (OVI) results in a smaller mass of liquid fuel stored on the port wall and
more liquid fuel transport directly into the cylinder than CVI [17]. The intake airflow causes the
injected liquid fuel distribution in the port to be significantly different for OVI injection than CVI.
This difference in port fuel distribution has a direct impact on the liquid fuel distribution in the
cylinder.
A semi-quantitative assessment of the liquid fuel transport into the cylinder can be made
using the injector probe location, estimated airflow velocity profiles, port geometry, and
combustion chamber design. We recognize three different types of physical transport of fuel into
the combustion chamber, 1) fuel which enters the chamber as vapor, 2) droplets which are
suspended in the airflow and may evaporate during transport or impinge on combustion chamber
surfaces, and 3) a liquid fuel film or puddle, that enters into the combustion chamber as a liquid
film attached to the port and combustion chamber walls. Through the visualization study of the
fuel injector probes we found that the injector probe deposits the liquid fuel in the form of a
puddle on the intake port wall. This puddle is then transported into the cylinder by one of the
three basic transport methods described above.
By examining the intake velocity profile from a similarly designed 4-valve engine we can
estimate the velocity profiles in the Nissan engine during the intake process. Kim and Lee did a
CFD analysis of the intake process in a 4-valve head with a head and port geometry very similar to
the Nissan engine used in this study. Their CFD simulation was at WOT with one of the intake
valves closed. While this CFD simulation is for a higher intake flow rate, the intake flow profile is
not expected to be significantly different than our 0.5 Bar MAP experiments. However, intake
105
valve flow profile experiments on a 4-valve head, by Lou and Daneshyar, found that the velocity
profiles around the intake valve were independent of flow rate; the velocity magnitudes where
found to be almost proportional to flow rate [29].
5.2 Comparison of Liquid Fuel Probes to Standard PFI
To interpret the results of the injector fuel probe experiments a comparison between the
fuel probe locations must be made to the liquid fuel distribution in a standard port fuel injection
configuration. Meyer [3] found that during OVI the dominant transport mechanism of liquid fuel
into the cylinder was fuel droplets transported directly from the injector.
The fuel droplets
entrained in the airflow will enter the cylinder along the same flow path as the intake airflow.
Most of the fuel droplets entrained during OVI will enter the cylinder over the top of the valves
where the airflow is highest [3]. OVI results in most of the liquid fuel entering the cylinder at a
location that is similar to location "A" of the injector probe. During CVI the liquid fuel will be
deposited along the port walls based on injector targeting. The standard injector on the Nissan
engine is targeted to deposit liquid fuel on the lower port wall and outer edge of the intake valves.
A semi-quantitative comparison can be made between the average of the probe "B" and "C"
locations with the standard Nissan injector targeting during CVI.
5.3 Conclusions
The purpose of this work was to develop a fundamentally based description of liquid fuel
transport into the engine cylinder of a port fuel injected, gasoline fueled, SI engine, and to develop
a method of quantifying the liquid fuel contribution to HC emissions. A novel liquid fuel injector
probe was developed which deposits controlled amounts of liquid fuel onto three precise
circumferential locations around the intake port, just upstream of the intake valve seat.
By
controlling the amount and location of the liquid fuel injected at the valve seat, during open valve
injection, and measuring the engine-out hydrocarbon (HC) emissions an estimate of the
contribution to these emissions, due to liquid fuel entering the cylinder, can be made.
conclusions from these experiments are as follows:
106
The
1)
The impact of the liquid fuel injected at the valve seat was studied by fueling the engine with
pre-vaporized Indolene plus different amounts of liquid Indolene injected through the fuel
probes.
These experiments indicated that the engine-out HC's increase linearly as the
proportion of liquid fuel, relative to pre-vaporized fuel, is increased. The rate of increase in
HC's is different for each probe location, and is different for closed valve injection (CVI) vs.
open valve injection (OVI).
2) The blowback of hot residual gases during valve overlap redistributes the liquid fuel in the
intake port and aids in the fuel air mixing process for CVI.
The liquid fuel probe
experiments show a similar difference between OVI and CVI hydrocarbons at all three probe
locations. The blowback flow velocities around the valve seat are approximately equal at all
probe locations. At time of injection, the liquid fuel film distribution is similar at all three
probe locations. Therefore, the blowback has the same effect at all three probe locations
resulting in the constant difference between OVI and CVI, at each probe location.
3) The location of liquid fuel around the valve seat has a significant impact on engine-out
hydrocarbon emissions. All the fuel probes were located just upstream of the valve seat and
injected liquid fuel film onto the intake port wall in a similar manner. These experiments
indicated that for all amounts of liquid fuel injected at the valve seat, the fuel delivered
closest to the exhaust valve (probe "A") resulted in the highest engine-out HC's, while the
location farthest from the exhaust valve (probe "C") had the lowest HC's.
A method of interpretation of the experimental probe data was developed, based on
the mass of the HC's, assuming a linear relationship between the mass of liquid fuel injected
at the valve seat and the total HC's. The constants of this linear relationship YWq and Yvap
were estimated by a linear regression analysis of the experimental data. Examining the ratio
of Y/iq/Yvap we find that during OVI the liquid fuel fractions of probe locations A, B, and C
compared to the pre-vaporized liquid fuel fractions are 7 times, 5 times, and 3 times the
value of the pre-vaporized liquid fuel fraction. During CVI, the liquid fuel fraction of probe
locations A, B, and C are 5 times, 2 times, and 1.3 times the pre-vaporized fuel fraction.
This analysis indicates that a given amount of liquid fuel entering the cylinder as liquid at
107
probe location "A" (closest to the exhaust valve) will result in 7 times the engine out HC's of
an equal amount of fuel entering the cylinder as vaporized fuel. While a given amount of
liquid fuel entering the cylinder as a liquid at probe location "C" (farthest from the exhaust
valve), will result in 3 times the HC's of an equal amount of vaporized fuel.
4) The difference in HC's at each probe location is due the subsequent distribution of liquid
fuel in the cylinder. Stanglmaier [14] studied the effect of in-cylinder wall wetting location
on HC emissions using a specially designed spark plug mounted injector probe.
His
experimental results suggest that the differences in the HC for the different cylinder wall
wetting locations can be explained in terms of the physical location of the liquid fuel relative
to the exhaust valves. He explains, liquid fuel deposited on the cylinder wall is scraped into
the piston top land during the compression stroke and then laid back on the cylinder wall
during expansion as a liquid fuel film. Any of this fuel film that survives the combustion
and post-flame oxidation process will be scraped into the roll-up vortex during the exhaust
stroke. The portions of the roll-up vortex closest to the exhaust valves have a higher chance
of being exhausted, and contributing to engine-out HC's, than the portions of the vortex
farther from the exhaust valves.
Therefore, liquid fuel entering the exhaust side of the
cylinder will have a higher probability of being exhausted during the exhaust stroke. Liquid
fuel deposited on the intake side of the cylinder has a lower probability of being exhausted
[14].
CFD data indicates that liquid fuel deposited on the intake port wall closest to the
exhaust valve (probe "A") will enter onto the exhaust side of the cylinder head (onto the
exhaust valves, between the exhaust valves, and into the exhaust valve crevices) and onto the
exhaust side of the cylinder liner resulting in the highest HC's. Liquid fuel deposited farthest
from the exhaust valve (probe "C") will enter onto the intake side of the cylinder and result
in lower HC's relative to probe "A". These experiments qualitatively agree with the incylinder liquid fuel injection experiments done by Stanglmaier and Matthew [14][15].
5) The overall physical conclusions of these experiments are that liquid fuel entering the
cylinder increases the engine-out HC's, relative to that fuel entering in vaporized form. The
engine-out HC's increase proportionally to the amount of liquid fuel entering the cylinder.
108
By modifying liquid fuel distribution in the intake port and the subsequent distribution of
liquid fuel in the cylinder, the impact of liquid fuel entering the cylinder on HC's can be
minimized. Liquid fuel entering the cylinder results in less of an increase in HC's if the
liquid fuel is deposited farther away from the exhaust valves.
The redistribution and
vaporization of liquid fuel due to exhaust backflow during CVI further reduces the impact of
a given amount of liquid fuel deposited in the intake port. A given amount of fuel deposited
at probe location "C" (farthest from the exhaust valve) with CVI results in only 1.3 times the
HC's of an equal amount of vaporized fuel, compared to 7 times at the location "A" (closest
the exhaust valve) with OVI of liquid fuel. In short, fully vaporizing the fuel results in the
lowest HC's. When liquid fuel does reach the valve seat region in the intake port, the lowest
HC's are achieved by closed valve injection, with the liquid fuel at the valve seat
predominantly on the side of the intake port farthest from the exhaust valves.
109
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110
Figure 5.1 CFD Simulation of 4-valve head, at WOT with one intake
valve closed, 1500 rpm.[28] Nissan head configuration overlay.
Figure 5.2 Relative liquid fuel transport mechanisms for probe locations
"A" and "C"
111
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112
Fuel from
Probe
C
A(
Liquid Fuel
Fuel enterin
Cylinder
Vaporized
Fuel at Start
of Combustion
.DVaporized
Vaporized
Fi-el Film,
Liquid
Vaporized
Fuel DL
Liquid
Probe
A
B
C
Fuel Drops
H
M
IL
Fuel Film
L
M
H
Vaporization Relative HC
H
L
M
M
IH
|L
Figure 5.3. Relative comparison of liquid fuel transport mechanisms
for different probe locations
113
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114
REFERENCES
[1]
Heywood, J.B., Internal Combustion Engine Fundamentals, McGraw-Hill Book Co., pp. 4245, 1988.
[2]
Automotive Handbook (4 th edition), Robert Bosch GmbH, pp.458-470, 1996.
[3]
Meyer, R., "Liquid Fuel Transport into the Cylinder in a Spark Ignition Engines," PhD.
Thesis Department of Mechanical Engineering, MIT, 1998.
[4]
Cheng, W.K., D. Hamrin, J.B. Heywood, S. Hochgreb, M. Min, and M. Norris, "An
Overview of Hydrocarbon Emissions Mechanisms in Spark-Ignition Engines," SAE Paper
932708, 1993.
[5]
Min, K, W.K. Cheng, and J.B. Heywood, "The Effects of Crevices on the Engine-Out
Hydrocarbon Emissions in SI Engines," SAE 940306, 1994.
[6]
Sterlepper J., U. Spicher, and H. Ruhland, "Flame Propagation into Top Land Crevice and
hydrocarbon Emissions from a SI Engine During Engine Warm-Up," SAE Paper 932648.
1993.
[7]
Haidar, H.A., and J.B. Heywood, " Combustion Chamber Deposit Effects on Hydrocarbon
Emissions from a Spark-Ignition Engine," SAE 972887, 1997.
[8]
Linna, J., Malberg, H., Bennett, P.J., Palmer, P.J., Tian, T., and Cheng W.K., "Contribution
of Oil Layer Mechanism to the Hydrocarbon Emissions from Spark-Ignition Engines," SAE
paper 972892, 1997.
[9]
Kaiser, E.W., Seigel, W.O., and Russ, S.G., "Effect of Fuel dissolved in Crankcase Oil on
Engine-Out Hydrocarbon Emissions from a Spark-Ignited Engine," SAE 972891, 1997.
[10] Norris, M.G., and Hochgreb, S., "Extent of Oxidation of Hydrocarbons Desorbing from the
Lubricant Oil Layer in Spark-Ignition Engines," SAE 960069, 1993.
[11] Parks, J., Armfield, J., Storey, J., Barber, and T., Wachter, E., "In Situ Measurement of Fuel
Absorption into the Cylinder Wall Oil Film During Engine Cold Start," SAE 981054, 1998.
[12] Meernik, P.R., Alkidas, A. C., "Impact of Exhaust Valve Leakage on Engine-Out
Hydrocarbons," SAE 932752, 1993.
[13] Hoard, J., Moilanen, P., "Exhaust Valve Seat Leakage," SAE 971638, 1997.
115
[14] Stanglmaier, R. H., Li, J., and Matthew, R. D., " The Effect of In-Cylinder Wall Wetting
Location on the HC Emissions from SI Engines," SAE 1999-01-0502, 1999.
[15] Li, J., Matthew, R. D., Stanglmaier, R. H., Roberts, C. E., and Anderso, R. W., " Further
Experiments on the Effects of In-Cylinder Wall Wetting on HC Emissions from Direct
Injection Gasoline Engines," SAE 1999-01-3661, 1999.
[16] Min, K., "The Effects of Crevices on the Engine-Out Hydrocarbon Emissions in Spark
Ignition Engines," Ph.D Thesis Department of Mechanical Engineering, MIT, 1993.
[17] Imatake, N., Saito, K., Morishima, S., Kudo, S., and Ohhata, A., "Quantitative Analysis of
Fuel Behavior in Port-Injection Gasoline Engines," SAE 971639, 1997.
[18] Kaiser, E. W., Siegl, W. 0., Lawson, G. P., Connolly, F. T., Cramer, C. F., Dobbins, K. L.,
Roth, P. W., Smokovitz, M., " Effect of Fuel Preparation on Cold-Start Hydrocarbon
Emissions from a Spark-Ignited Engine," SAE 961957, 1997.
[19] Shayler, P. J., Davies, M. T., " Audit of Fuel Utilization During the Warm-up of SI
Engines," SAE 971656, 1997.
[20] Fry, M., Nightingale, C., Richardson, S., " High-Speed Photography and Image Analysis
Techniques Applied to Study Droplet Motion within the Porting and Cylinder of a 4-valve SI
Engine," SAE 952525, 1995.
[21] Witze, P. 0., Green, R. M., "LIF and Flame-Emission Imaging of Liquid Fuel Films and
Pool Fires in an SI Engine During a Simulated Cold Start," SAE 970866, 1997.
[22] Yang, J., Kaiser, E. W., Siegl, W. 0., and Anderson, R. W., " Effects of Port-Injection
Timing and Fuel Droplet Size on Total Speciated Exhaust Hydrocarbon Emissions," SAE
930711, 1993.
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