Diesel Engine Modeling in WAVE
2004
Brian Feldman
The Pennsylvania State University
Schreyer Honors College
College of Engineering
Diesel Engine Modeling in WAVE
A Thesis in
Mechanical Engineering
by
Brian David Feldman
©2004 Brian David Feldman
Submitted in Partial Fulfillment
of the Requirements
for the Degree of
Bachelor of Science
May 2004
We approve the thesis of Brian David Feldman.
Date of Signature
________________________________________
Daniel C. Haworth
Associate Professor of Mechanical Engineering
Thesis Co-Advisor
______________________
________________________________________
Domenic A. Santavicca
Professor of Mechanical Engineering
Honors Advisor
______________________
Abstract
The FutureTruck team’s main goals this year are to improve the emissions output
and fuel economy of their hybrid electric vehicle through a combination of engine
modifications and aftertreatments. Several engine modification techniques,
including intake throttling, thermal throttling, EGR (exhaust gas recirculation),
and variable displacement diesel were modeled using a diesel engine model
developed in Ricardo’s WAVE software. Data from the baseline engine model
were compared to data obtained on an engine dyno to ensure an accurate baseline
model. The results from the model can be used to predict general trends in engine
performance characteristics if certain modifications were to be made to the actual
engine. According to data obtained from the model, intake throttling and cooled
EGR appear to be very promising from a fuel consumption and exhaust
aftertreatment perspective.
iii
Table of Contents
List of Figures
V
Acknowledgements
X
Chapter 1 – Goals of Powertrain Modifications
1
Chapter 2 – Engine Control Strategies
8
Intake Throttling
Thermal Intake Throttling
Variable Displacement Diesel
Exhaust Gas Recirculation
Exhaust Backpressure Increase
Timing Retardation
Chapter 3 – Engine Model
25
Chapter 4 – Engine Lab
28
Chapter 5 – Results
41
Intake Throttling
Thermal Intake Throttling
Variable Displacement Diesel
Exhaust Gas Recirculation
Hot EGR
Cooled EGR
Exhaust Backpressure Increase
Timing Retardation
Chapter 6 – Research Summary and Conclusions
94
Chapter 7 – Future Work
98
iv
List of Figures
Figure 2.1: Visual representation of intake throttling
9
Figure 2.2: Visual representation of thermal throttling
12
Figure 2.3: Visual representation of variable displacement diesel
15
Figure 2.4: Visual representation of hot EGR
18
Figure 2.5: Visual representation of cooled EGR
18
Figure 2.6: Timing retardation effect on combustion temperatures
23
Figure 3.1: Image of Engine Model in WAVE
26
Figure 4.1: Peak BHP vs. RPM, Lab vs. Wave vs. Rated
31
Figure 4.2: Minimum BSFC vs. RPM, Lab vs. Wave vs. Rated
32
Figure 4.3: Baseline model EGT (K) vs. BHP at 1000 rpm
32
Figure 4.4: Baseline model BSFC (kg/kwhr) vs. BHP at 1000 rpm
33
Figure 4.5: Baseline model BSFC (kg/kwhr) vs. BHP at 1000 rpm
33
Figure 4.6: Baseline model EGT (K) vs. BHP at 2000 rpm
34
Figure 4.7: Baseline model BSFC (kg/kwhr) vs. BHP at 2000 rpm
34
Figure 4.8: Baseline model BSFC (kg/kwhr) vs. BHP at 2000 rpm
35
Figure 4.9: Baseline model EGT (K) vs. BHP at 3000 rpm
35
Figure 4.10: Baseline model BSFC (kg/kwhr) vs. BHP at 3000 rpm
36
Figure 4.11: Baseline model EGT (K) vs. BHP at 4000 rpm
36
Figure 4.12: Baseline model BSFC (kg/kwhr) vs. BHP at 4000 rpm
37
Figure 4.13: Baseline Model Fuel Consumption Map
39
Figure 4.14: Baseline Model EGT map
40
Figure 5.1: Intake Throttled EGT (K) vs. BHP at 700 rpm
42
Figure 5.2: Intake Throttled BSFC (kg/kwhr) vs. BHP at 1000 rpm
43
v
Figure 5.3: Intake Throttled EGT (K) vs. BHP at 1000 rpm
43
Figure 5.4: Intake Throttled BSFC (kg/kwhr) vs. BHP at 1000 rpm
44
Figure 5.5: Intake Throttled EGT (K) vs. BHP at 2000 rpm
44
Figure 5.6: Intake Throttled BSFC (kg/kwhr) vs. BHP at 2000 rpm
45
Figure 5.7: Intake Throttled EGT (K) vs. BHP at 3000 rpm
45
Figure 5.8: Intake Throttled BSFC (kg/kwhr) vs. BHP at 3000 rpm
46
Figure 5.9: Intake Throttled EGT (K) vs. BHP at 4000 rpm
46
Figure 5.10: Intake Throttled BSFC (kg/kwhr) vs. BHP at 4000 rpm
47
Figure 5.11: Intake Throttled Fuel Consumption Map
47
Figure 5.12: Intake Throttled EGT Map
48
Figure 5.13: Thermal Throttled EGT (K) vs. BHP at 1000 rpm
50
Figure 5.14: Thermal Throttled BSFC (kg/kwhr) vs. BHP at 1000 rpm
50
Figure 5.15: Thermal Throttled EGT (K) vs. BHP at 2000 rpm
51
Figure 5.16: Thermal Throttled BSFC (kg/kwhr) vs. BHP at 2000 rpm
51
Figure 5.17: Thermal Throttled EGT (K) vs. BHP at 3000 rpm
52
Figure 5.18: Thermal Throttled BSFC (kg/kwhr) vs. BHP at 3000 rpm
52
Figure 5.19: Thermal Throttled EGT (K) vs. BHP at 4000 rpm
53
Figure 5.20: Thermal Throttled BSFC (kg/kwhr) vs. BHP at 4000 rpm
53
Figure 5.21: Variable Displacement Diesel EGT (K) vs. BHP at 700 rpm
56
Figure 5.22: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 700 rpm
57
Figure 5.23: Variable Displacement Diesel EGT (K) vs. BHP at 1000 rpm
57
Figure 5.24: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 1000 rpm 58
Figure 5.25: Variable Displacement Diesel EGT (K) vs. BHP at 1300 rpm
58
Figure 5.26: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 1300 rpm 59
Figure 5.27: Variable Displacement Diesel EGT (K) vs. BHP at 2000 rpm
59
Figure 5.28: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 2000 rpm 60
vi
Figure 5.29: Variable Displacement Diesel EGT (K) vs. BHP at 3000 rpm
60
Figure 5.30: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 3000 rpm 61
Figure 5.31: Variable Displacement Diesel EGT (K) vs. BHP at 4000 rpm
61
Figure 5.32: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 4000 rpm 62
Figure 5.33: Variable Displacement Diesel Fuel Consumption Map
63
Figure 5.34: Variable Displacement Diesel EGT Map
63
Figure 5.35: Hot EGR EGT (K) vs. BHP at 1000 rpm
66
Figure 5.36: Hot EGR BSFC (kg/kwhr) vs. BHP at 1000 rpm
67
Figure 5.37: Hot EGR EGT (K) vs. BHP at 2000 rpm
67
Figure 5.38: Hot EGR BSFC (kg/kwhr) vs. BHP at 2000 rpm
68
Figure 5.39: Hot EGR EGT (K) vs. BHP at 3000 rpm
68
Figure 5.40: Hot EGR BSFC (kg/kwhr) vs. BHP at 3000 rpm
69
Figure 5.41: Hot EGR EGT (K) vs. BHP at 4000 rpm
69
Figure 5.42: Hot EGR BSFC (kg/kwhr) vs. BHP at 4000 rpm
70
Figure 5.43: Hot EGR % of Exhaust Gas in Intake Map
70
Figure 5.44: Cooled EGR EGT (K) vs. BHP at 1000 rpm
72
Figure 5.45: Cooled EGR BSFC (kg/kwhr) vs. BHP at 1000 rpm
72
Figure 5.46: Cooled EGR EGT (K) vs. BHP at 2000 rpm
73
Figure 5.47: Cooled EGR BSFC (kg/kwhr) vs. BHP at 2000 rpm
73
Figure 5.48: Cooled EGR EGT (K) vs. BHP at 3000 rpm
74
Figure 5.49: Cooled EGR BSFC (kg/kwhr) vs. BHP at 3000 rpm
74
Figure 5.50: Cooled EGR EGT (K) vs. BHP at 4000 rpm
75
Figure 5.51: Cooled EGR BSFC (kg/kwhr) vs. BHP at 4000 rpm
75
Figure 5.52: Cooled EGR % of Exhaust Gas in Intake Map
76
Figure 5.53: Cooled EGR PPM NO Map
76
vii
Figure 5.54: Baseline PPM NO Map
77
Figure 5.55: Cooled EGR BSNO2 Map
77
Figure 5.56: Baseline BSNO2 Map
78
Figure 5.57: 700 mbar max backpressure EGT (K) vs. BHP at 1000 rpm
81
Figure 5.58: 700 mbar max backpressure BSFC (kg/kwhr) vs. BHP at 1000 rpm
81
Figure 5.59: 700 mbar max backpressure EGT (K) vs. BHP at 2000 rpm
82
Figure 5.60: 700 mbar max backpressure BSFC (kg/kwhr) vs. BHP at 2000 rpm
82
Figure 5.61: 700 mbar max backpressure EGT (K) vs. BHP at 3000 rpm
83
Figure 5.62: 700 mbar max backpressure BSFC (kg/kwhr) vs. BHP at 3000 rpm
83
Figure 5.63: 700 mbar max backpressure EGT (K) vs. BHP at 4000 rpm
84
Figure 5.64: 700 mbar max backpressure BSFC (kg/kwhr) vs. BHP at 4000 rpm
84
Figure 5.65: 700 mbar max Backpressure Exhaust System Pressure Map
85
Figure 5.66: 700 mbar max Backpressure EGT Map
85
Figure 5.67: Timing Retarded 15 deg EGT (K) vs. BHP at 1000 rpm
87
Figure 5.68: Timing Retarded 15 deg BSFC (kg/kwhr) vs. BHP at 1000 rpm
87
Figure 5.69: Timing Retarded 15 deg EGT (K) vs. BHP at 2000 rpm
88
Figure 5.70: Timing Retarded 15 deg BSFC (kg/kwhr) vs. BHP at 2000 rpm
88
Figure 5.71: Timing Retarded 15 deg EGT (K) vs. BHP at 3000 rpm
89
Figure 5.72: Timing Retarded 15 deg BSFC (kg/kwhr) vs. BHP at 3000 rpm
89
Figure 5.73: Timing Retarded 15 deg EGT (K) vs. BHP at 4000 rpm
90
Figure 5.74: Timing Retarded 15 deg BSFC (kg/kwhr) vs. BHP at 4000 rpm
90
Figure 5.75: Timing Retarded 15 deg Fuel Consumption Map
91
Figure 5.76: Timing Retarded 15 deg EGT Map
91
Figure 5.77: Timing Retarded 15 deg PPM NO Map
92
Figure 5.78: Baseline PPM NO Map
92
Figure 5.79: Timing Retarded 15 deg BSNO2 Map
93
viii
Figure 5.80: Baseline BSNO2 Map
93
ix
Acknowledgements
I would like to thank the following:
Dan Haworth for his support and guidance throughout the research project
Fawzan Al-Sharif of Ricardo for his time and training in WAVE modeling
software
The Penn State FutureTruck Emissions Team for setting up the engine on the
dyno and collecting data
The Penn State FutureTruck Team for revealing the need for research like this to
further their own understanding
Brian Feldman
x
Chapter 1 – Goals of Powertrain Modifications
Overall Goals of Powertrain Modifications
The goals of modeling the 2.5L Detroit Diesel engine used for this thesis are to
investigate the effects of various simple and inexpensive engine modifications
upon engine performance and emissions. Many engine modifications involve
compromises and tradeoffs. A reduction in one type of emissions may accompany
an increase in several other types of emissions plus an increase in fuel
consumption. The main purpose in modeling the modifications is to quantify the
effects of each one. This will allow vehicle designers to make more educated
decisions about how to best utilize the engine for minimum emissions and
maximum performance and economy.
One group that stands to reap immediate benefits from the knowledge gained
through this research is the Penn State FutureTruck Hybrid Electric Vehicle
Team. The FutureTruck team is using the same 2.5L Detroit Diesel engine that
the simulation model is derived from. In addition to having detailed fuel
consumption and emissions data and being able to determine which projects to
pursue for maximum benefit and having the data to back up those decisions, the
team will have a detailed and fairly accurate model upon which future testing can
be quickly and easily performed.
1
The Need for Engine Modeling
Modeling an engine through software is one of the least expensive and quickest
methods of obtaining reasonably accurate data based on reasonably accurate
assumptions. Operating conditions and modifications that would require
significant amounts of time and money to test can be modeled to obtain
information that is accurate enough to make informed decisions and determine
major effects. Information that could not be obtained through conventional
methods can be obtained from a model as well. Modeling the 2.5L Detroit Diesel
engine will help the FutureTruck team to make quick and informed decisions
about which modifications to the engine will help them to best achieve their goals.
This research will develop a reasonably accurate model of the 2.5L Detroit Diesel
that will be useful for further research into engine modifications and control
strategies on this particular engine as well as similar engines.
Top Priorities for FutureTruck Project
The FutureTruck project aims to modify a stock vehicle to achieve 25% better
fuel economy and lower emissions while retaining the capabilities and features of
a stock 2002 Ford Explorer. [FutureTruck 2004] Fifteen schools across the
nation participate in this annual competition. The improvements in fuel economy
and emissions are achieved through the development of an appropriate hybrid
electric powertrain for the vehicle. Penn State’s strategy incorporates a diesel
2
engine in conjunction with an electric motor and battery pack. The diesel engine
is chosen because it has a higher thermal efficiency than a comparable gasoline
engine. The tradeoff in choosing a diesel engine is that emissions are much
tougher to control due to particulate matter formation and the inability to employ
a conventional catalytic converter due to excess oxygen in the exhaust stream.
One major issue with aftertreatment of exhaust is that the treatment components
work best within a certain range of exhaust gas temperatures. Research into
methods of increasing the exhaust gas temperatures is important to the Penn State
FutureTruck team because during the 2002 competition, the diesel particulate
filter being used was clogged due to exhaust gas temperatures that were too low to
regenerate the filter, as a result of extended idling during the competition. Penn
State would benefit the most at competition by concentrating on reducing criteria
pollutants, including PM (particulate matter), CO (carbon monoxide), HC
(hydrocarbons), and NOx (oxides of nitrogen).
Reduction of Emissions Through Engine Control
Penn State chose a turbocharged 103 kW 2.5L Detroit Diesel engine for use in the
truck because they have extensive experience with this particular engine and it
was one of the few diesel engines that was available at the time it was selected
with a power rating that met the requirements for the hybrid powertrain design.
Unfortunately this engine is equipped with a closed Bosch ignition and engine
control system so that modifying the ignition system is almost impossible.
3
Computer controls and sensors used are calibrated specifically for the stock
engine and thus extensive modifications to the engine itself would likely not result
in improvements to fuel efficiency or emissions. Without extensive modifications
to the engine, emissions control can still be achieved through trying to run the
engine near certain load and speed ranges by use of the hybrid powertrain,
exhaust aftertreatment, and exhaust gas recirculation.
Ideally to reduce engine-out NOx emissions, peak combustion temperatures need
to be reduced. A significant reduction in NOx emissions could perhaps come from
a diesel engine based on the Atkinson cycle, in which the compression and
expansion ratio are variable. If fuel could be injected into the cylinder over a
longer period of time, this would slightly decrease efficiency but would greatly
reduce the temperature spike characteristically observed in Otto cycle and Diesel
cycle combustion. The Atkinson cycle engine also offers higher efficiency than an
Otto or Diesel cycle engine due to the adjustable compression and expansion
ratio, which could offset any efficiency penalties due to delayed timing. If NOx
formation is prevented well enough during combustion, aftertreatment is
unnecessary. This is especially important since NOx is one of the hardest
emissions to treat. The Atkinson cycle reduces available power and torque, but if
used in conjunction with an electric motor as with a hybrid vehicle the effects can
be minimized. The 2004 Toyota Prius and 2005 Ford Escape HEV (hybrid
electric vehicle) both use Atkinson cycle gasoline engines in their hybrid
4
powertrains. [Toyota 2004, Ford Motor Company 2004] A similar concept for a
diesel should be investigated.
Reduction of Emissions Through Aftertreatment
NOx, CO, HC, and PM can all be reduced through the use of commercially
available aftertreatment products. Penn State currently uses a urea SCR (selective
catalytic reduction) system to treat NOx emissions. The use of a urea SCR
aftertreatment system for NOx reduction such as the Penn State FutureTruck team
uses may be impractical for the typical consumer because it is costly to implement
and maintain, requires precise calibration and control, and requires frequent
refills. [National Laboratory for the Environment 2004] Since the consumer will
notice no degradation in operation of their vehicle when such a system is not
functioning, they will have no incentive to maintain the system and may not even
be aware that it is not working properly. CO and HC emissions can be reduced by
an oxidation catalyst, while particulate matter emissions are most effectively
treated by a diesel particulate filter. Diesel particulate filters require the exhaust
gas temperature to be above a certain point to regenerate. [Brewbaker 2002]
Unfortunately, emissions aftertreatment products all take up space, add weight,
and add backpressure to the engine. Penn State must be careful of the amount of
backpressure the exhaust aftertreatment adds because higher levels of
backpressure reduce fuel efficiency and power output from the engine. Since Penn
State has determined that aftertreatment is the most effective method of emissions
5
control for this particular application, all efforts need to be taken to ensure that the
aftertreatment functions as effectively as possible. This includes appropriately
sizing aftertreatment components, placing them in appropriate locations, and
ensuring that the exhaust gas temperatures are within acceptable ranges.
Particulate matter, hydrocarbons, and smoke may form from combustion,
especially if combustion is optimized for low NOx production due to the typical
NOx – PM tradeoff for diesels. A diesel particulate filter with a low regeneration
temperature would greatly reduce all three of these pollutants and if operating
properly would not add an enormous amount of backpressure. It would also
function as a muffler, eliminating the need for a separate sound reduction device.
Particulate filters do need to be cleaned periodically to remove ash and other
deposits. This service would likely only need to be performed after every 30,000
– 50,000 miles, which means that it could be done at the same time as other
major servicing is done on the vehicle, such as replacing tires or brakes.
Reduction of Fuel Consumption
Penn State reduces the fuel consumption of the FutureTruck by replacing a
gasoline engine with a diesel for higher thermal efficiency and employing a
hybrid electric powertrain so that the engine can be run near its most efficient
operating conditions and braking losses can be partially recovered. Since it has
been determined that major engine modifications are infeasible, the most effective
6
remaining ways to reduce fuel consumption include lightweighting the vehicle,
reducing accessory loads where possible, refining the hybrid powertrain control
algorithm to run the engine near its most efficient operating conditions, ensuring
that the engine intake is not restricted, and keeping the exhaust backpressure as
low as possible. The FutureTruck competition emphasizes mainly low-speed startand-stop driving, so modifying the aerodynamics of the vehicle to reduce fuel
consumption would not be very effective.
7
Chapter 2 – Engine Control Strategies
Intake Throttling
Theory of Intake Throttling to Increase EGT for Emissions Aftertreatment
Restricting the airflow into the engine is one relatively simple way to increase the
EGT (exhaust gas temperature) that Penn State could implement on the
FutureTruck without too much trouble. Throttling reduces the amount of air
available to the engine for combustion. The same amount of fuel is burned with
less air, resulting in a higher fuel-air ratio than would result from normal
operation at a given operational point. For this experiment the fuel-air ratio will
be kept at a constant 0.045 over all operating points tested. Since less excess
oxygen is present during combustion, the net energy released by combustion heats
less matter than it would if more excess oxygen were present, resulting in higher
temperatures. [Mayer 2003] This could be accomplished with the use of a simple
throttle plate, which is commonly available due to its use on virtually every
production gasoline engine. Intake throttling is likely more viable than electrically
heating the exhaust or using a heated diesel particulate filter or catalyst due to
conversion losses for electricity production, high power requirements for electric
heating methods, increased loads on the electrical system, and difficulty in finding
and implementing electric exhaust heating products. Figure 2.1 demonstrates the
8
differences between an engine operating normally and an engine operating in a
throttled manner. Note that during normal operation the throttle plate is open,
whereas it is nearly closed during throttled operation.
Figure 2.1: Visual representation of intake throttling
Basic Strategy for use of Intake Throttling
Intake throttling would only be necessary when the EGT is below the required
range for exhaust aftertreatment, such as during periods of operation at or near
idle. Since EGT can be measured or easily calculated based on engine model
maps it would be simple to determine when and to what degree the engine needs
to be throttled to keep the EGT in the appropriate range. By limiting use of intake
9
throttling to the periods in which it is necessary, adverse effects on fuel economy
and emissions are minimized. Reduction in peak power output of the engine using
this control strategy is not an issue because the engine would not be throttled
under normal and high-load operation, as the EGT would already be high enough
under these operating conditions. Penn State also uses insulation around the
exhaust system to keep the temperature of exhaust gases reaching aftertreatment
devices higher due to a reduction of heat lost through the piping.
Advantages and Disadvantages of Intake Throttling
Intake throttling is relatively easy to implement on the FutureTruck. Throttle
plates are small, light, inexpensive, and commonly available. Exhaust gas
temperatures can be easily measured or computed to determine the degree of
throttling necessary. Throttling is one of the less energy-intensive methods of
increasing exhaust gas temperature. It has no adverse effects on peak power
output and does not affect the engine at all over most of the operating range.
Intake throttling is also not critical to engine operation, so the loss of the ability to
throttle the intake would not prevent the FutureTruck from participating in
competition.
Intake throttling does increase fuel consumption at idle by a measurable amount
and may adversely affect engine out NOx emissions. Accelerator position input to
the engine may need to be modified to raise the fuel-air ratio to prevent the engine
10
from stalling when throttling is employed. Sensors such as the MAF (mass
airflow sensor) may need to be bypassed to prevent the engine from setting error
codes and shutting down.
Thermal Intake Throttling
Theory of Thermal Intake Throttling
By increasing the air temperature of the intake the amount of air entering the
cylinders is reduced because the density of air is reduced by increasing its
temperature at a fixed pressure. If the same amount of fuel is burned, the fuel-air
ratio would be higher than for a colder, denser charge of air, which means less
excess oxygen would be present, resulting in higher exhaust gas temperatures.
Also, the temperature of the exhaust will be higher partially due to the fact that
the temperature of the air at the beginning of combustion is higher. Figure 2.2
demonstrates the differences between an engine operating normally and an engine
operating in a thermally throttled manner.
11
Figure 2.2: Visual representation of thermal throttling
Basic Strategy for Use of Thermal Intake Throttling
Thermal intake throttling could be accomplished in several ways. A secondary air
intake could be located closer to the engine, radiator, or exhaust system to take in
warmer air when desired. An intercooler bypass could be designed to eliminate
the ability of the intercooler to cool intake air after passing through the
compressor. Another method of warming intake air would be to incorporate a heat
exchanger to take heat from the exhaust or engine coolant. By limiting use of
thermal intake throttling to the periods in which it is necessary, adverse effects on
fuel economy and emissions are minimized. Reduction in peak power output of
12
the engine using this control strategy is not an issue because the engine would not
be thermally throttled under normal and high-load operation, as the EGT would
already be high enough under these operating conditions.
Advantages and Disadvantages of Thermal Intake Throttling
Thermal intake throttling would not be critical for operation of the truck, and the
inability to use the system would not prevent the FutureTruck from competing.
Intake air temperature and EGT can be easily measured to determine the amount
of thermal throttling necessary. There would be no adverse effects on peak power
output of the engine as thermal throttling would only be employed near idle
conditions.
Thermal intake throttling would be tougher to implement than simple pressure
throttling because of the need for a secondary air intake in a warm location and a
method to mix warm and cold intake air in the correct proportions. Warm air for
thermal throttling would not be available immediately upon engine startup due to
the fact that the engine would be cold. If the intake air becomes too warm, the
engine may set an error code and shut down. The MAF and other sensors may
need to be bypassed to prevent this.
13
Variable Displacement Diesel
Theory of Using Variable Displacement in a Diesel Engine
To the knowledge of the FutureTruck team, the effects of variable displacement
have never been investigated in a diesel engine before. Major automakers are
starting to employ variable displacement in gasoline engines. Chrysler’s 5.7L
Hemi, which is being used in the 2005 Chrysler 300C and 2005 Dodge Magnum
is one such engine. [Chrysler 2004] Variable displacement involves modifying
the engine so that combustion occurs in only half of the cylinders in an engine.
Since the cylinders that would be firing would be running under much more load
than they would be if all of the cylinders were firing, the belief is that an engineout emissions reduction can be achieved and fuel consumption will decrease since
the engine normally operates more efficiently and produces lower emissions when
operating under heavier loads. It has also been suggested that the engine chosen
for the FutureTruck is moderately oversized and reducing effective displacement
through cylinder deactivation would offer the benefits of the economy of a
smaller engine with the power availability of a larger engine, and it would be
simpler to deactivate the cylinders than to install a smaller engine. Figure 2.3
demonstrates the engine operating in a variable displacement manner with two
cylinders turned off. In a production application the engine would likely be
designed initially to accommodate variable displacement. This model is being
used to determine if anything can be gained from a simpler implementation.
14
Figure 2.3: Visual representation of variable displacement diesel
Basic Strategy for use of Variable Displacement in a Diesel
Variable displacement would be used at or near idle operation to increase the load
factor on the firing cylinders with the hopes of reducing emissions and fuel
consumption. The engine must have the ability to run on all cylinders if necessary
for peak power output during events such as acceleration and towing at
competition. Engine load is easily monitored and controlled, which would make it
easy to determine when to activate and deactivate cylinders.
15
Advantages and Disadvantages of Variable Displacement in a Diesel
Deactivating cylinders on an engine produces a higher degree of torque pulsation
because the engine is getting two power strokes per revolution instead of four.
This can lead to significant vibrations since the engine is designed and balanced
to operate with four cylinders firing. Special balancing pendulums would be
necessary to counteract these vibrations. [Nester 2003] Deactivation of cylinder
injectors will set engine trouble codes and cause the engine to shut down entirely
unless the engine control module is fooled or overridden. Exhaust gas
temperatures are not increased at low loads because cold intake air is pumped
through deactivated cylinders directly into the exhaust. Sensors such as the MAF
may need to be bypassed because of the vast quantity of air coming in that would
not be used in combustion. Brake specific fuel consumption increases due to the
effects of running two cylinders without combustion occurring. It might also be
tough to make a smooth transition between running on two and four cylinders.
Exhaust Gas Recirculation
Theory of using Exhaust Gas Recirculation
16
Exhaust gas recirculation involves rerouting a fraction of exhaust gases from the
exhaust manifold to the air intake of the engine. The goal is to reduce engine out
NOx emissions by altering the combustion process. Recirculation of exhaust gas
has several effects on the combustion process. If the exhaust gas is not cooled
before being introduced into the intake, it will heat up the incoming air charge and
provide a thermal throttling effect. Adding exhaust gas to the intake also raises the
fuel-to-air ratio by lowering the concentration of oxygen. This has the effect of
lowering combustion temperatures by delaying the combustion. The heat capacity
of exhaust gas is also higher than that of ambient air, resulting in lower
combustion temperatures as well. Since NOx formation occurs due to the presence
of nitrogen and oxygen combined with high temperatures, the reduction of
temperatures in the combustion chamber lowers the amount of NOx produced.
[Kouremenos 2001] Figure 2.4 demonstrates the differences between an engine
operating normally and an engine operating with hot EGR, while figure 2.5 shows
an engine operating with cooled EGR.
17
Figure 2.4: Visual representation of hot EGR
Figure 2.5: Visual representation of cooled EGR
18
Basic Strategy for use of EGR
Exhaust gas recirculation can be employed at any time during engine operation to
lower NOx emissions. The recirculation of exhaust gases depends on a differential
pressure between the exhaust and intake. At high loads and speeds with a
turbocharged engine, the pressure differential is high enough to force a significant
quantity of exhaust gases to recirculate. [Van Nieuwstadt 2003] At lower loads
and operating points without a large enough pressure differential, intake throttling
can be used to lower the intake pressure and raise the pressure differential
between the exhaust and the intake. The exhaust gases can be introduced into the
intake at a high temperature or cooled.
Advantages and Disadvantages of EGR
Exhaust gas recirculation is fairly easy to utilize and the 2.5L Detroit Diesel, as
well as many other engines, are designed with a built-in EGR system which is
active in the stock configuration. An EGR system does not add a significant
amount of weight, cost, size, or complexity to an engine. Proper operation of the
EGR system is not critical to the engine’s ability to operate. Exhaust gas
recirculation can lower peak combustion temperatures, resulting in lower NOx
production and may even allow the engine to be tuned to run more efficiently
while meeting emissions criteria.
19
The main disadvantages of EGR are a reduction in peak power output, a possible
increase in brake specific fuel consumption at high speeds and loads, and the
potential for increased PM, CO and HC emissions. There are also durability issues
with EGR, especially in a diesel. Reduction of engine-out NOx emissions is not
critical to the FutureTruck team’s performance due to the fact that a urea SCR
system has been added to the truck specifically for the purpose of treating NOx
emissions. The FutureTruck team would gain the most benefit by deactivating the
EGR system to increase peak power at high engine speeds and loads.
Increasing Exhaust Backpressure
Theory of Increasing Exhaust Backpressure
Increasing exhaust backpressure is similar to intake throttling in that it results in
increased pumping work for the engine, resulting in an increase in fuel
consumption. By restricting the exhaust flow, extra work will be necessary to
overcome the restriction, which will result in higher pressures and temperatures
ahead of the restriction. Investigating and quantifying the effects of increasing
exhaust backpressure is important when designing an emissions aftertreatment
system because many aftertreatment devices, such as mufflers, catalysts, and
20
filters add backpressure to an exhaust system. If the effects of the additional
backpressure are not accounted for, exhaust gas temperatures, power output, and
fuel consumption may differ greatly from expectations. Exhaust backpressure can
also be intentionally added to the system by using a throttle plate in the exhaust if
desired.
Strategy of Increasing Exhaust Backpressure
Exhaust backpressure can be created by adding an adjustable restriction in the
exhaust stream, or, more conventionally, by adding emissions aftertreatment
devices such as mufflers, catalysts, and filters. By adding restrictions to the
exhaust stream the EGT will rise with added backpressure. The goal in
investigating backpressure is to ensure that fuel consumption, exhaust gas
temperatures, and power output are all maintained at acceptable levels.
Advantages and Disadvantages of Increasing Exhaust Backpressure
Adding exhaust aftertreatment devices to the exhaust stream can be incredibly
beneficial for the reduction of undesirable pollutants, such as CO, HC, PM, and
NOx. However, the sizing and usage of these aftertreatment devices and the
backpressure they create must be balanced against allowable fuel consumption,
EGT, and power requirements, especially at high loads and speeds. An
21
excessively high backpressure may also damage the engine or components in the
exhaust system by creating extremely high temperatures.
Timing Retardation
Theory of Timing Retardation
Timing retardation can be easily accomplished by modifying the engine control
unit to inject fuel later than it normally would so that combustion and the phasing
of heat released are delayed. By causing combustion and therefore heat release to
occur later in the power stroke, the amount of expansion the combustion gases
undergo is reduced. Less expansion means that less work is performed, resulting
in an increase in fuel consumption, reduction in power, and an increase in exhaust
gas temperatures. However, one of the appealing benefits of timing retardation is
that peak cylinder temperatures are reduced, which reduces the formation of NOx,
one of the toughest pollutants to treat in a diesel engine. [Kouremenos 2001]
Figure 2.6 is a representation which demonstrates the differences between an
engine operating normally and an engine operating in a timing retarded manner.
22
Figure 2.6: Timing retardation effect on combustion temperatures
Strategy of Timing Retardation
Timing retardation could be employed any time it is necessary to reduce engineout NOx emissions or increase the exhaust gas temperature. Exhaust gas
temperatures would primarily need to be increased at low engine loads. Engine
timing could be left unchanged at high engine speeds and loads to avoid a
reduction in peak power output and increases in fuel consumption at high power
levels.
23
Advantages and Disadvantages of Timing Retardation
Timing retardation is very easy to control and modify with an open engine control
unit. It requires no additional modifications to an engine and no extra expense.
Timing retardation can be adjusted so that it is employed when it provides the
most benefit. Peak power and fuel consumption levels can remain unchanged if
the timing is only retarded at lower speeds and loads. Less engine-out NOx is
created. Exhaust gas temperatures increase with timing retardation, although the
effect is more pronounced at higher engine loads. Unfortunately, fuel
consumption increases and power output decreases when the timing is retarded.
24
Chapter 3 – Engine Model
Ricardo’s WAVE Modeling Software
WAVE by Ricardo (http://www.ricardo.com) was used to create the 2.5L Detroit
Diesel engine model. WAVE is a comprehensive engine modeling package used
in the automotive industry to develop engines and determine performance levels
and other characteristics before an engine is actually built. Characteristics such as
valves, injectors, compression ratios, fuel type, and the heat release rate due to
combustion can all be modified to determine their effects. WAVE employs onedimensional time-dependent computational fluid dynamics to model an engine. A
simulated engine run to create an engine map consists of modeling the engine at
many different speed and load points, with enough iterations at each point to
reach approximate convergence of all factors.
Model Overview
Figure 3.1 is an image of the engine model constructed in WAVE.
25
Figure 3.1: Image of Engine Model in WAVE
The red area indicates ambient air and the air intake system. Air moves along to
the orange area, where it passes through the compressor of the turbocharger and
then through the intercooler. It continues along through the yellow area, which is
the intake manifold. The green area indicates the four cylinders and fuel injectors.
Exhaust is pumped into the exhaust manifold, marked in blue. Exhaust then
passes through the turbine of the turbocharger, indicated in purple. The remaining
exhaust aftertreatment components and end of the tailpipe are circled in brown.
The grey circle marks the exhaust gas recirculation system, in which some
26
exhaust gas is taken from the exhaust manifold, cooled via an EGR cooler, and
then introduced back into the intake manifold.
There are many input parameters in the model that affect engine performance.
These parameters include but are not limited to: duct length, duct temperatures,
thermal conductivities, valve timing, crankshaft speed, cylinder dimensions,
compression ratio, fuel-air ratio, injection timing, combustion timing, combustion
heat release profile, turbocharger wastegate opening, EGR valve opening,
frictional losses, pressure drops in the ductwork, turbocharger performance maps,
fuel properties, and ambient conditions.
Output data available from the model include but is not limited to: pressure traces,
temperature traces, turbocharger speed, NOx, PM, and HC emissions, power
output, fuel consumption, volumetric efficiency, scavenging efficiency, exhaust
gas temperatures, and percent exhaust gases in intake air.
27
Chapter 4 – Engine Dyno in Lab
The engine that was used to calibrate the model was set up in the Penn State
Academic Activities Building by the FutureTruck Emissions team and testing was
conducted during Spring 2004. A blend of biodiesel, B35, was used for testing, as
this was the same specification as the fuel to be used at competition. For the initial
testing and baselining no emissions aftertreatment devices were installed and
emissions were not measured. The EGR was also disabled. Due to time
constraints, only one baseline test matrix could be run before the model needed to
be calibrated against it. Fortunately, the data points obtained appear to be
reasonably consistent and are close to expected values. Further testing on the
engine in the lab will include measuring engine out emissions and the
effectiveness of the emissions aftertreatments. A recently acquired open engine
control unit will also allow the team to fine-tune the engine specifically for the
hybrid-electric vehicle powertrain.
Comparison of Model to 2.5L Detroit Diesel
The WAVE engine model used to model the actual 2.5L Detroit Diesel engine is
based off a model created by Sara Inman for her master’s thesis. [Inman 2002]
After much modification using data obtained in the lab, it now compares quite
well with the engine in the lab in terms of characteristics such as power output,
exhaust gas temperatures, and fuel consumption. Unfortunately, due to the
28
unavailability of complete information about the engine, certain assumptions
about operating parameters had to be made, such as injection and combustion
timing, turbocharger and compressor maps, turbocharger wastegate settings, and
EGR settings. Emissions output was not concentrated on in great detail because
the FutureTruck was being designed to run with several different emissions
aftertreatment devices to mitigate the effects of engine-out emissions and specific
information about the shape and properties of the combustion chamber in the
engine were unknown. The accuracy of engine-out emissions projections from the
model would therefore be extremely rough guesses at best and not of great
importance due to the aftertreatments.
Comparing engine models with real test data and published data [Detroit Diesel
2004] using data points obtained at the peak power output alone is not an
entirely accurate method of verification for several reasons. One of the main
reasons is that there are so many factors in the model that can be adjusted, such as
wastegate opening and fuel-air ratio, which cannot be easily measured or
determined from manufacturer’s specifications. Another reason is that test
conditions and methods in the dyno lab may be slightly different than the
manufacturer’s test conditions and methods. Of greater importance in the engine
model is a reasonable correlation between parameters such as fuel consumption
and exhaust gas temperatures at different loads and speeds, which can be easily
measured and provide more useful information for calibration of the vehicle
systems over the entire range of operation. Figure 4.1 compares the peak power
29
output of the engine in the lab, the rated specifications, and the WAVE baseline
model. Figure 4.2 compares the minimum BSFC (brake specific fuel
consumption) for the engine in the lab, the published specifications, and the
WAVE baseline model. Figures 4.3 through 4.12 compare BSFC and EGT data
obtained from the lab and model baseline over the entire operating range of the
engine.
One of the main differences between the model, rated power output, and engine in
the lab was that the engine in the lab was fueled with a biodiesel blend of B35,
whereas the model and manufacturer used regular diesel fuel. Biodiesel has been
known to increase fuel consumption slightly and lower exhaust temperatures a
small amount. [Fedak 2003] Since the blend of biodiesel used in the engine lab
was only a 35% blend of biodiesel, these effects of using B35 instead of regular
diesel were assumed to be very minor and would not have a significant impact on
the engine calibration or the trends that the model was used to examine.
Another difference between the model, rated power output, and engine in the lab
is a difference in backpressures. The backpressure on the engine in the lab was
not measured directly but is likely to be very low because there were no flow
restricting devices in the exhaust system. The backpressure on the baseline model
had no more than 110 mbar of maximum backpressure, whereas the rated
specifications allow for a backpressure of 250 mbar.
30
During calibration between the engine on the dyno and the WAVE model, EGR
was disabled and there were no exhaust aftertreatment devices. This eliminated
the need to account for two variables that could make calibrating the WAVE
model much more difficult. However, the potential of the model to include the
effects of both EGR and emissions aftertreatment devices still remains.
BHP - Lab/Wave/Rated
160
140
Power (bhp)
120
100
Lab
Wave
Rated
80
60
40
20
0
1000
1500
2000
2500
3000
3500
4000
RPM
Figure 4.1: Peak BHP vs. RPM, Lab vs. Wave vs. Rated
31
BSFC (kg/kwh) - Lab/Wave/Rated
0.300
BSFC (kg/kwhr)
0.250
0.200
Lab
0.150
Wave
Rated
0.100
0.050
0.000
1000
1500
2000
2500
3000
3500
4000
RPM
Figure 4.2: Minimum BSFC vs. RPM, Lab vs. Wave vs. Rated
1000 RPM EGT
1000
EGT (K)
900
800
Lab
700
Wave
600
500
400
0
5
10
15
20
25
30
Power (bhp)
Figure 4.3: Baseline model EGT (K) vs. BHP at 1000 rpm
32
1000 RPM BSFC
BSFC (kg/kwhr)
1.200
1.000
0.800
Lab
0.600
Wave
0.400
0.200
0.000
0
5
10
15
20
25
30
Power (bhp)
Figure 4.4: Baseline model BSFC (kg/kwhr) vs. BHP at 1000 rpm
1000 RPM BSFC
BSFC (kg/kwhr)
0.400
0.300
Lab
0.200
Wave
0.100
0.000
0
5
10
15
20
25
30
Power (bhp)
Figure 4.5: Baseline model BSFC (kg/kwhr) vs. BHP at 1000 rpm
33
2000 RPM EGT
1000
EGT (K)
900
800
Lab
700
Wave
600
500
400
0.00
20.00
40.00
60.00
80.00
100.00
Power (bhp)
Figure 4.6: Baseline model EGT (K) vs. BHP at 2000 rpm
2000 RPM BSFC
BSFC (kg/kwhr)
1.000
0.800
0.600
Lab
0.400
Wave
0.200
0.000
0
20
40
60
80
100
Power (bhp)
Figure 4.7: Baseline model BSFC (kg/kwhr) vs. BHP at 2000 rpm
34
2000 RPM BSFC
BSFC (kg/kwhr)
0.300
0.250
0.200
Lab
0.150
Wave
0.100
0.050
0.000
0
20
40
60
80
100
Power (bhp)
Figure 4.8: Baseline model BSFC (kg/kwhr) vs. BHP at 2000 rpm
3000 RPM EGT
1000
EGT (K)
900
800
Lab
700
Wave
600
500
400
0
50
100
150
Power (bhp)
Figure 4.9: Baseline model EGT (K) vs. BHP at 3000 rpm
35
3000 RPM BSFC
BSFC (kg/kwhr)
0.500
0.400
0.300
Lab
0.200
Wave
0.100
0.000
0
50
100
150
Power (bhp)
Figure 4.10: Baseline model BSFC (kg/kwhr) vs. BHP at 3000 rpm
4000 RPM EGT
1000
EGT (K)
900
800
Lab
700
Wave
600
500
400
0
50
100
150
200
Power (bhp)
Figure 4.11: Baseline model EGT (K) vs. BHP at 4000 rpm
36
BSFC (kg/kwhr)
4000 RPM BSFC
0.400
0.350
0.300
0.250
0.200
0.150
0.100
0.050
0.000
Lab
Wave
0
50
100
150
200
Power (bhp)
Figure 4.12: Baseline model BSFC (kg/kwhr) vs. BHP at 4000 rpm
Figures 4.3 to 4.12 show that the exhaust gas temperatures obtained from the
model correlate quite well with the values obtained in the lab at engine speeds
above 1500 rpm. At nearly all operating points at speeds of 2000 rpm and above,
the exhaust gas temperature difference between the lab and model is within 50 K
and also, more importantly, the trends are consistent. At 1500 rpm and below the
engine control unit for the engine in the lab was thought to be operating in a
different manner than normal to control operating conditions at idle, and the full
effects of these differences were unknown. Further testing in the lab would be
necessary to determine the accuracy of the EGT values gathered as there appear to
be a few minor anomalies in the data, such as an unexplained rise in the EGT in
the mid-range power output at 2000 rpm. Differences in the combustion heat
release profile, heat transfer rates, and start-of-combustion timing may be some of
the reasons that the EGT values obtained from the model differ from those in the
lab.
37
At engine loads above approximately 20% of full power for any given speed,
brake specific fuel consumption (BSFC) values obtained from the model and the
lab appear to correlate quite well and appear to be within 3-5% at most points and
trends appear consistent. Below 20% load the values differ, sometimes by a
significant amount. This might be due to the fact that the dyno measures power
output after driveshaft losses, whereas the model does not factor driveshaft losses
into its power output. Another reason the BSFC values obtained in the lab at very
low power outputs may have such high variation is that the engine is clearly
operating very inefficiently in this range, and very small changes in the fueling
rate might cause large changes in power output since the BSFC varies greatly
relative to power output, making it difficult to obtain accurate values. Further
testing in the lab would be necessary to determine the accuracy of the BSFC
values gathered as there appear to be a few minor anomalies in the data, such as
random high and low BSFC points on the 1000 rpm and 4000 rpm operating
conditions that do not appear to be in line with other points gathered. Differences
in the combustion heat release profile, heat transfer rates, and start-of-combustion
timing may be some of the reasons that the BSFC values obtained from the model
differ slightly from those in the lab. Figures 4.13 and 4.14 show maps of BSFC
and EGT, respectively, over the entire operating range of the model baseline
engine.
38
Figure 4.13: Baseline Model Fuel Consumption Map
39
Figure 4.14: Baseline Model EGT map
40
Chapter 5 – Results
Intake Throttling
The results obtained from the intake throttling model match expectations. To
model an intake throttle, an orifice with a selectable diameter was used ahead of
the compressor to create a pressure drop. The fuel injectors were set to provide a
constant fuel-air ratio of 0.045 for all cases in the intake throttling test. These
results show that a constant EGT can be obtained across the entire load range at
any speed, although the maximum attainable EGT is dependent upon speed. The
EGT can be adjusted as desired at any speed by varying the amount of throttling
and the fuel-air ratio.
Effects on Engine Out Emissions and Fuel Consumption
Intake throttling is a restriction on the intake of the engine, which increases
pumping losses, and thus brake specific fuel consumption increases when
throttling is used. The amount of extra fuel consumed and increase in exhaust gas
temperature is determined by the degree of throttling employed. Engine-out NOx
increases but PM is reduced due to the higher combustion temperatures. The
impact on emissions and fuel consumption is minimized because intake throttling
is only necessary when the EGT is too low for aftertreatment devices to be
effective, such as during extended periods of operation at or near idle. During
41
normal driving conditions intake throttling is not necessary, and thus fuel
economy and emissions would not be adversely affected during most of the
driving cycle. Figures 5.1 through 5.10 compare BSFC and EGT data obtained
from the model baseline and intake throttled model over the entire operating range
of the engine. These graphs show that EGT can be maintained at a high level
regardless of load by throttling, even though fuel consumption increases slightly.
Figures 5.11 and 5.12 provide maps of BSFC and EGT, respectively, near idle
from the intake throttled model.
700 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Intake Throttle Idle
600
500
400
0
5
10
15
Power (bhp)
Figure 5.1: Intake Throttled EGT (K) vs. BHP at 700 rpm
42
BSFC (kg/kwhr)
700 RPM BSFC
0.700
0.600
0.500
0.400
0.300
0.200
0.100
0.000
Wave Baseline
Intake Throttle Idle
0
5
10
15
Power (bhp)
Figure 5.2: Intake Throttled BSFC (kg/kwhr) vs. BHP at 1000 rpm
1000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Intake Throttle
600
500
400
0
5
10
15
20
25
Power (bhp)
Figure 5.3: Intake Throttled EGT (K) vs. BHP at 1000 rpm
43
BSFC (kg/kwhr)
1000 RPM BSFC
0.700
0.600
0.500
0.400
0.300
0.200
0.100
0.000
Wave Baseline
Intake Throttle
0
5
10
15
20
25
Power (bhp)
Figure 5.4: Intake Throttled BSFC (kg/kwhr) vs. BHP at 1000 rpm
2000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Intake Throttle
600
500
400
0
20
40
60
80
100
Power (bhp)
Figure 5.5: Intake Throttled EGT (K) vs. BHP at 2000 rpm
44
BSFC (kg/kwhr)
2000 RPM BSFC
0.700
0.600
0.500
0.400
Wave Baseline
0.300
0.200
0.100
0.000
Intake Throttle
0
20
40
60
80
100
Power (bhp)
Figure 5.6: Intake Throttled BSFC (kg/kwhr) vs. BHP at 2000 rpm
3000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Intake Throttle
600
500
400
0
50
100
150
Power (bhp)
Figure 5.7: Intake Throttled EGT (K) vs. BHP at 3000 rpm
45
3000 RPM BSFC
BSFC (kg/kwhr)
1.000
0.800
0.600
Wave Baseline
0.400
Intake Throttle
0.200
0.000
0
50
100
150
Power (bhp)
Figure 5.8: Intake Throttled BSFC (kg/kwhr) vs. BHP at 3000 rpm
4000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Intake Throttle
600
500
400
0
50
100
150
200
Power (bhp)
Figure 5.9: Intake Throttled EGT (K) vs. BHP at 4000 rpm
46
BSFC (kg/kwhr)
4000 RPM BSFC
0.700
0.600
0.500
0.400
Wave Baseline
0.300
0.200
Intake Throttle
0.100
0.000
0
50
100
150
200
Power (bhp)
Figure 5.10: Intake Throttled BSFC (kg/kwhr) vs. BHP at 4000 rpm
Figure 5.11: Intake Throttled Fuel Consumption Map
47
Figure 5.12: Intake Throttled EGT Map
48
Thermal Intake Throttling
Thermal intake throttling was modeled by increasing the ambient temperature and
temperatures of all components in the intake system by approximately 50K. The
ambient temperature was changed to 350K from 298K. In practice, thermal
throttling could be accomplished by blowing warm air across the intercooler,
bypassing the intercooler, or using a heat exchanger between the intake and
exhaust. The results obtained match the expected results.
Effects on Engine Out Emissions and Fuel Consumption
Thermal intake throttling reduces the density of the air entering the engine, which
reduces the power output of the engine without reducing the intake pressure.
Since less air enters the cylinders the fuel-air ratio is effectively increased without
a corresponding increase in power output. The amount of extra fuel consumed and
increase in exhaust gas temperature is determined by the degree of thermal
throttling employed. Engine-out NOx increases but PM is reduced due to the
higher combustion temperatures. The impact on emissions and fuel consumption
is minimized because thermal intake throttling is only necessary when the EGT is
too low for aftertreatment devices to be effective, such as during extended periods
of operation at or near idle. During normal driving conditions intake throttling is
not necessary, and thus fuel economy and emissions would not be adversely
affected during most of the driving cycle. Figures 5.13 through 5.20 compare
49
BSFC and EGT data obtained from the model baseline and thermal throttled
model over the entire operating range of the engine. These show that at all
operating points thermal throttling increases EGT and slightly increases BSFC.
1000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Thermal Throttle
600
500
400
0
5
10
15
20
25
Power (bhp)
Figure 5.13: Thermal Throttled EGT (K) vs. BHP at 1000 rpm
1000 RPM BSFC
BSFC (kg/kwhr)
0.400
0.300
Wave Baseline
0.200
Thermal Throttle
0.100
0.000
0
5
10
15
20
25
Power (bhp)
Figure 5.14: Thermal Throttled BSFC (kg/kwhr) vs. BHP at 1000 rpm
50
2000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Thermal Throttle
600
500
400
0
20
40
60
80
100
Power (bhp)
Figure 5.15: Thermal Throttled EGT (K) vs. BHP at 2000 rpm
2000 RPM BSFC
BSFC (kg/kwhr)
0.300
0.250
0.200
Wave Baseline
0.150
Thermal Throttle
0.100
0.050
0.000
0
20
40
60
80
100
Power (bhp)
Figure 5.16: Thermal Throttled BSFC (kg/kwhr) vs. BHP at 2000 rpm
51
3000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Thermal Throttle
600
500
400
0
50
100
150
Power (bhp)
Figure 5.17: Thermal Throttled EGT (K) vs. BHP at 3000 rpm
3000 RPM BSFC
BSFC (kg/kwhr)
0.600
0.500
0.400
Wave Baseline
0.300
Thermal Throttle
0.200
0.100
0.000
0
50
100
150
Power (bhp)
Figure 5.18: Thermal Throttled BSFC (kg/kwhr) vs. BHP at 3000 rpm
52
4000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Thermal Throttle
600
500
400
0
50
100
150
200
Power (bhp)
Figure 5.19: Thermal Throttled EGT (K) vs. BHP at 4000 rpm
4000 RPM BSFC
BSFC (kg/kwhr)
1.000
0.800
0.600
Wave Baseline
0.400
Thermal Throttle
0.200
0.000
0
50
100
150
200
Power (bhp)
Figure 5.20: Thermal Throttled BSFC (kg/kwhr) vs. BHP at 4000 rpm
53
Variable Displacement Diesel
Variable Displacement Diesel was modeled by removing the fuel injectors from
cylinders 2 and 3, which fire opposite each other. The firing sequence of this
engine is 1-3-4-2. The valvetrain was left fully operational because it would be
impractical to shut down the valves for any cylinders unless the engine was
originally designed with this in mind. In some cases a higher EGT can be
obtained by shutting off two of the cylinders, but in all cases fuel consumption is
drastically increased. Peak power output is dramatically decreased to
approximately 1/3 of the peak power obtainable when all cylinders are firing. If
this strategy were to be employed, provisions would need to be made to run on all
cylinders as running in two-cylinder mode is incredibly inefficient, vastly reduces
available power and torque, and nasty vibrations might be encountered. It would
make much more sense to downsize to a more appropriately sized engine to
reduce weight and increase efficiency if it is decided that the engine in use is
oversized or investigate the possibility of controlling air flow in addition to fuel
flow.
Effects on Engine Out Emissions and Fuel Consumption
Unfortunately it is impractical to stop the valvetrain on the two non-firing
cylinders without major engine modifications and engineering effort. One idea for
implementing variable displacement involved removing the glow plugs for the
54
two non-firing cylinders and replacing them with electronically actuated valves to
reduce pumping losses when running in two-cylinder mode. However, this would
limit the ability of the engine to start properly when cold, and it could severely
alter characteristics such as combustion chamber geometry, cylinder leakage, and
compression ratio. Due to these constraints, it was decided that the easiest way to
implement variable displacement would be to cut off fuel injector pulses to the
two non-firing cylinders and devise a method to fool the engine control module
into sensing that the injectors were receiving pulses to prevent an engine fault
from registering.
Since the valvetrain would still be running, intake air would be flowing to all of
the cylinders, and exhaust gas from all the cylinders would flow into the exhaust
manifold. Because combustion would not be occurring in two cylinders a large
amount of cold unburned air would enter the exhaust stream. For this reason the
use of variable displacement to raise the exhaust gas temperature is not as
effective as it would be if the valvetrain for the two non-firing cylinders could be
stopped.
During two-cylinder operation, the effects of friction and pressure differentials
between the exhaust and intake in combination with the running valvetrain rob the
engine of power, resulting in significantly higher brake specific fuel consumption.
Running a four cylinder diesel engine on two cylinders in this manner would not
increase exhaust gas temperatures at all at very low loads and speeds. Since
55
exhaust gas temperatures need to be raised most at idle and two-cylinder
operation cannot be used at high loads due to limits on power output, this method
of operation offers no tangible benefits in its current configuration unless it
reduces emissions. Figures 5.21 through 5.32 compare BSFC and EGT data
obtained from the model baseline and variable displacement model over the entire
operating range of the engine. These show that variable displacement significantly
increases fuel consumption and only provides a small increase in EGT at low
speeds with light-to-moderate loads. Figures 5.33 and 5.34 provide maps of BSFC
and EGT, respectively, from the variable displacement diesel model.
700 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
VDD Idle
600
500
400
0
5
10
15
Power (bhp)
Figure 5.21: Variable Displacement Diesel EGT (K) vs. BHP at 700 rpm
56
700 RPM BSFC
BSFC (kg/kwhr)
0.600
0.500
0.400
Wave Baseline
0.300
VDD Idle
0.200
0.100
0.000
0
5
10
15
Power (bhp)
Figure 5.22: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 700
rpm
1000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
VDD
600
500
400
0
5
10
15
20
25
Power (bhp)
Figure 5.23: Variable Displacement Diesel EGT (K) vs. BHP at 1000 rpm
57
1000 RPM BSFC
BSFC (kg/kwhr)
1.000
0.800
0.600
Wave Baseline
0.400
VDD
0.200
0.000
0
5
10
15
20
25
Power (bhp)
Figure 5.24: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 1000
rpm
1300 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
VDD Idle
600
500
400
0
10
20
30
40
Power (bhp)
Figure 5.25: Variable Displacement Diesel EGT (K) vs. BHP at 1300 rpm
58
1300 RPM BSFC
BSFC (kg/kwhr)
0.600
0.500
0.400
Wave Baseline
0.300
VDD Idle
0.200
0.100
0.000
0
10
20
30
40
Power (bhp)
Figure 5.26: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 1300
rpm
2000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
VDD
600
500
400
0
20
40
60
80
100
Power (bhp)
Figure 5.27: Variable Displacement Diesel EGT (K) vs. BHP at 2000 rpm
59
2000 RPM BSFC
BSFC (kg/kwhr)
1.000
0.800
0.600
Wave Baseline
0.400
VDD
0.200
0.000
0
20
40
60
80
100
Power (bhp)
Figure 5.28: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 2000
rpm
3000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
VDD
600
500
400
0
50
100
150
Power (bhp)
Figure 5.29: Variable Displacement Diesel EGT (K) vs. BHP at 3000 rpm
60
3000 RPM BSFC
BSFC (kg/kwhr)
1.200
1.000
0.800
Wave Baseline
0.600
VDD
0.400
0.200
0.000
0
50
100
150
Power (bhp)
Figure 5.30: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 3000
rpm
4000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
VDD
600
500
400
0
50
100
150
200
Power (bhp)
Figure 5.31: Variable Displacement Diesel EGT (K) vs. BHP at 4000 rpm
61
BSFC (kg/kwhr)
4000 RPM BSFC
1.600
1.400
1.200
1.000
0.800
0.600
0.400
0.200
0.000
Wave Baseline
VDD
0
50
100
150
200
Power (bhp)
Figure 5.32: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 4000
rpm
62
Figure 5.33: Variable Displacement Diesel Fuel Consumption Map
Figure 5.34: Variable Displacement Diesel EGT Map
63
Exhaust Gas Recirculation
Hot EGR was modeled by removing the EGR intercooler in the model and
opening an orifice to a diameter of 1 cm to let gases flow through the EGR system
from the exhaust manifold (before the turbine) to the intake manifold (after the
compressor) due to a pressure gradient. As can be seen in the EGR map, Figure
5.43, appreciable rates of EGR only occur above 2000 rpm. This may be due to
inaccuracies in the model calibration because assumptions were made on factors
such as turbocharger and wastegate settings and timings. The drop in EGT at
higher loads at 1000 rpm as seen in Figure 5.35 may be due to flow reversal in the
EGR path. [Yang 2003] Pressure differences in the system may be such that cool
intake air was flowing into the exhaust, cooling it down. [Van Nieuwstadt 2003],
[Jacobs 2003] The maximum amount of EGR in this case occurs at the higher
loads at 3000 and 4000 rpm, and the expected effects of increased EGT and
slightly increased fuel consumption are present. However, retuning the engine
properly to accommodate EGR might lower the increase in exhaust gas
temperatures and reduce fuel consumption while also reducing NOx.
Effects on Engine out Emissions and Fuel Consumption
Exhaust gas recirculation has been shown to reduce NOx emissions due to the
reduction of combustion temperatures. Combustion deterioration at higher levels
of exhaust gas recirculation can lead to increased CO and HC emissions.
64
However, exhaust gas recirculation can lead to higher brake specific fuel
consumption because of decreased engine output due to lower combustion
chamber temperatures and increased pumping losses associated with maintaining
the necessary pressure differential between the exhaust and intake to achieve
exhaust gas flow. The reduction in fuel efficiency is related to the amount of
exhaust gas recirculation employed and the degree of pumping losses. Presumably
the negative impact on efficiency could be minimized by altering the combustion
process to account for the effects of exhaust gas recirculation and designing the
turbocharging setup for minimal pumping losses. [Jacobs 2003] Uncooled EGR
increases the charge temperature, which assists NOx formation. Thus it is
preferable to cool the EGR if the main goal is NOx reduction. [Yang 2003] In the
WAVE model, the amount of exhaust gas recirculated reached a maximum of 12
to 15% of intake air at high loads and speeds because the highest pressure
differential between exhaust and intake was obtained at these operating
conditions. By investigating the results from those particular operating points, it
can be seen that heated EGR raises the exhaust gas temperature, while cooled
EGR has very little effect on exhaust gas temperature, but lowers fuel
consumption slightly compared to operation without EGR. More work is needed
to definitively determine the true effects of a properly calibrated EGR system as
factors like injection timing and start of combustion may need to be modified to
produce peak efficiency and emissions reduction when using EGR. For this
experiment all factors were kept the same as during operation without EGR. Also,
the beneficial effects of EGR may be enhanced by increasing the amount of
65
exhaust gas recirculated at lower loads and speeds. With this model the effects of
EGR are only evident at high loads and speeds because sufficient amounts of
EGR only occur under these conditions. Figures 5.35 through 5.42 compare BSFC
and EGT data obtained from the model baseline and hot EGR model over the
entire operating range of the engine. These show that hot EGR increases the EGT
if the exhaust gas is flowing to the intake, and cools the EGT if the flow is
reversed. The effects on fuel economy appear to be minimal at most points.
Figure 5.43 provides a map showing the amount of exhaust gases recirculated into
the intake over the operational range of the engine occurring in the hot EGR
model. Note the large area where 0% of the composition of the intake gas is
exhaust gas. This indicates either no flow of exhaust into the intake or flow of
intake air directly into the exhaust through the EGR system.
Hot EGR
1000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
EGR Hot
600
500
400
0
5
10
15
20
25
Power (bhp)
Figure 5.35: Hot EGR EGT (K) vs. BHP at 1000 rpm
66
1000 RPM BSFC
BSFC (kg/kwhr)
0.600
0.500
0.400
Wave Baseline
0.300
EGR Hot
0.200
0.100
0.000
0
5
10
15
20
25
Power (bhp)
Figure 5.36: Hot EGR BSFC (kg/kwhr) vs. BHP at 1000 rpm
2000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
EGR Hot
600
500
400
0
50
100
150
Power (bhp)
Figure 5.37: Hot EGR EGT (K) vs. BHP at 2000 rpm
67
2000 RPM BSFC
BSFC (kg/kwhr)
0.600
0.500
0.400
Wave Baseline
0.300
EGR Hot
0.200
0.100
0.000
0
50
100
150
Power (bhp)
Figure 5.38: Hot EGR BSFC (kg/kwhr) vs. BHP at 2000 rpm
3000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
EGR Hot
600
500
400
0
50
100
150
Power (bhp)
Figure 5.39: Hot EGR EGT (K) vs. BHP at 3000 rpm
68
3000 RPM BSFC
BSFC (kg/kwhr)
0.600
0.500
0.400
Wave Baseline
0.300
EGR Hot
0.200
0.100
0.000
0
50
100
150
Power (bhp)
Figure 5.40: Hot EGR BSFC (kg/kwhr) vs. BHP at 3000 rpm
4000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
EGR Hot
600
500
400
0
50
100
150
200
Power (bhp)
Figure 5.41: Hot EGR EGT (K) vs. BHP at 4000 rpm
69
BSFC (kg/kwhr)
4000 RPM BSFC
0.800
0.700
0.600
0.500
0.400
0.300
0.200
0.100
0.000
Wave Baseline
EGR Hot
0
50
100
150
200
Power (bhp)
Figure 5.42: Hot EGR BSFC (kg/kwhr) vs. BHP at 4000 rpm
Figure 5.43: Hot EGR % of Exhaust Gas in Intake Map
70
Cooled EGR
Cooled EGR was modeled by opening an orifice to a diameter of 1 cm to let gases
flow through the EGR system from the exhaust manifold (before the turbine) to
the intake manifold (after the compressor) due to a pressure gradient. As can be
seen in the EGR map, Figure 5.52, appreciable rates of EGR only occur above
2000 rpm. This may be due to inaccuracies in the model calibration because
assumptions were made on factors such as turbocharger and wastegate settings
and timings. The drop in EGT at higher loads at 1000 rpm as seen in Figure 5.44
may be due to flow reversal in the EGR path. [Yang 2003] Pressure differences
in the system may be such that cool intake air was flowing into the exhaust,
cooling it down. [Van Nieuwstadt 2003], [Jacobs 2003] The maximum amount
of EGR in this case occurs at the higher loads at 3000 and 4000 rpm, and the
expected effects of increased EGT are present. The increase in EGT with cooled
EGR is less than the increase with hot EGR. However, it can be seen that fuel
consumption is actually reduced due to cooled EGR at 3000 and 4000 rpm. With
proper engine tuning fuel consumption with cooled EGR might be able to be
reduced even more, providing the benefits of reduced fuel consumption and
reduced NOx emissions. Figures 5.44 through 5.51 compare BSFC and EGT data
obtained from the model baseline and cooled EGR model over the entire
operating range of the engine. Fuel consumption appears to be lowered at high
speeds with high amounts of EGR occurring. Figure 5.52 provides a map showing
the amount of exhaust gases recirculated into the intake over the operational range
71
of the engine occurring in the cooled EGR model. Note the large area where 0%
of the composition of the intake gas is exhaust gas. This indicates either no flow
of exhaust into the intake or flow of intake air directly into the exhaust through
the EGR system. Figures 5.53 through 5.56 provide maps comparing PPMNO
(parts per million NO) and BSNO2 (brake specific NO2) emissions from the
baseline model and cooled EGR model.
1000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
EGR Cooled
600
500
400
0
5
10
15
20
25
Power (bhp)
Figure 5.44: Cooled EGR EGT (K) vs. BHP at 1000 rpm
1000 RPM BSFC
BSFC (kg/kwhr)
0.600
0.500
0.400
Wave Baseline
0.300
EGR Cooled
0.200
0.100
0.000
0
5
10
15
20
25
Power (bhp)
Figure 5.45: Cooled EGR BSFC (kg/kwhr) vs. BHP at 1000 rpm
72
2000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
EGR Cooled
600
500
400
0
20
40
60
80
100
Power (bhp)
Figure 5.46: Cooled EGR EGT (K) vs. BHP at 2000 rpm
2000 RPM BSFC
BSFC (kg/kwhr)
0.600
0.500
0.400
Wave Baseline
0.300
EGR Cooled
0.200
0.100
0.000
0
20
40
60
80
100
Power (bhp)
Figure 5.47: Cooled EGR BSFC (kg/kwhr) vs. BHP at 2000 rpm
73
3000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
EGR Cooled
600
500
400
0
50
100
150
Power (bhp)
Figure 5.48: Cooled EGR EGT (K) vs. BHP at 3000 rpm
3000 RPM BSFC
BSFC (kg/kwhr)
0.600
0.500
0.400
Wave Baseline
0.300
EGR Cooled
0.200
0.100
0.000
0
50
100
150
Power (bhp)
Figure 5.49: Cooled EGR BSFC (kg/kwhr) vs. BHP at 3000 rpm
74
4000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
EGR Cooled
600
500
400
0
50
100
150
200
Power (bhp)
Figure 5.50: Cooled EGR EGT (K) vs. BHP at 4000 rpm
BSFC (kg/kwhr)
4000 RPM BSFC
0.800
0.700
0.600
0.500
0.400
0.300
0.200
0.100
0.000
Wave Baseline
EGR Cooled
0
50
100
150
200
Power (bhp)
Figure 5.51: Cooled EGR BSFC (kg/kwhr) vs. BHP at 4000 rpm
75
Figure 5.52: Cooled EGR % of Exhaust Gas in Intake Map
Figure 5.53: Cooled EGR PPM NO Map
76
Figure 5.54: Baseline PPM NO Map
Figure 5.55: Cooled EGR BSNO2 Map
77
Figure 5.56: Baseline BSNO2 Map
78
Exhaust Backpressure Increase
Increasing the exhaust backpressure was accomplished in the model by including
ductwork in the exhaust system representing a passive emissions aftertreatment
device such as a muffler, catalyst, or filter. The ductwork consisted of many tiny
tubes, designed to increase the surface area and therefore provide friction and
resistance to flow on the moving air. This type of device will produce a varying
amount of backpressure with different amounts of flow, as seen in the map. As
expected, peak power output is greatly reduced, exhaust gas temperatures increase
in all instances, and fuel consumption rises. The maximum amount of
backpressure generated by the passive device in this experiment, 700 mbar, was
chosen to both represent an exhaust system with several aftertreatment devices in
a series configuration and to effectively show the effects of a backpressure well
above the 250 mbar used for the published engine performance ratings.
Effects of Increasing Exhaust Backpressure
Increases in exhaust backpressure result in higher EGT, higher fuel consumption,
and lower power output, as anything that increases the backpressure in an exhaust
system is a flow restriction, which increases the pumping work performed by the
engine. The amount of backpressure in an exhaust system with aftertreatment
devices varies with engine load and speed because the friction and pressure drop
across the devices is flow dependent. More backpressure is obtained at higher
79
exhaust flow rates. Therefore exhaust gas temperatures will increase most due to
backpressure at times when the exhaust is already the hottest, making the addition
of passive flow restriction devices in the exhaust stream a poor choice for
increasing EGT at low speeds and loads, which are the periods when EGT is
lowest and needs to be increased. Passive flow restriction devices can also greatly
reduce peak power output and increase fuel consumption because the most
backpressure is obtained at high speeds and loads when the most power is being
produced. Backpressure could also be added intentionally and controlled through
the use of a throttling device in the exhaust if desired. Figures 5.57 through 5.64
compare BSFC and EGT data obtained from the model baseline and increased
backpressure model over the entire operating range of the engine. Figures 5.65
and 5.66 provide maps showing the amount of exhaust backpressure and EGT,
respectively, over the operational range of the engine occurring in the
backpressure model. It can be clearly seen that fuel consumption increases over
the entire operating range. Note that backpressure generally increases with
exhaust flow rate and EGT rise is dependent on the amount of backpressure.
80
1000 RPM EGT
1000
EGT (K)
900
Wave Baseline
800
700
700 mbar
backpressure
600
500
400
0
5
10
15
20
25
Power (bhp)
Figure 5.57: 700 mbar max backpressure EGT (K) vs. BHP at 1000 rpm
BSFC (kg/kwhr)
1000 RPM BSFC
0.700
0.600
0.500
0.400
0.300
0.200
0.100
0.000
Wave Baseline
700 mbar
backpressure
0
5
10
15
20
25
Power (bhp)
Figure 5.58: 700 mbar max backpressure BSFC (kg/kwhr) vs. BHP at 1000
rpm
81
2000 RPM EGT
1000
EGT (K)
900
Wave Baseline
800
700
700 mbar
backpressure
600
500
400
0
20
40
60
80
100
Power (bhp)
Figure 5.59: 700 mbar max backpressure EGT (K) vs. BHP at 2000 rpm
BSFC (kg/kwhr)
2000 RPM BSFC
0.350
0.300
0.250
0.200
0.150
0.100
Wave Baseline
700 mbar
backpressure
0.050
0.000
0
20
40
60
80
100
Power (bhp)
Figure 5.60: 700 mbar max backpressure BSFC (kg/kwhr) vs. BHP at 2000
rpm
82
3000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
700 mbar
backpressure
600
500
400
0
50
100
150
Power (bhp)
Figure 5.61: 700 mbar max backpressure EGT (K) vs. BHP at 3000 rpm
3000 RPM BSFC
BSFC (kg/kwhr)
0.500
0.400
Wave Baseline
0.300
700 mbar
backpressure
0.200
0.100
0.000
0
50
100
150
Power (bhp)
Figure 5.62: 700 mbar max backpressure BSFC (kg/kwhr) vs. BHP at 3000
rpm
83
4000 RPM EGT
1000
EGT (K)
900
Wave Baseline
800
700
700 mbar
backpressure
600
500
400
0
50
100
150
200
Power (bhp)
Figure 5.63: 700 mbar max backpressure EGT (K) vs. BHP at 4000 rpm
4000 RPM BSFC
BSFC (kg/kwhr)
0.500
0.400
Wave Std
0.300
700 mbar
backpressure
0.200
0.100
0.000
0
50
100
150
200
Power (bhp)
Figure 5.64: 700 mbar max backpressure BSFC (kg/kwhr) vs. BHP at 4000
rpm
84
Figure 5.65: 700 mbar max Backpressure Exhaust System Pressure Map
Figure 5.66: 700 mbar max Backpressure EGT Map
85
Timing Retardation
Timing retardation was accomplished by delaying the injection and start of
combustion by 15 degrees. As expected, exhaust gas temperatures and fuel
consumption increased, quite drastically in some cases. A significant amount of
NOx reduction is also evident. The amount of timing retardation or advance could
be adjusted to provide the desired effects.
Effects of Timing Retardation
Timing retardation increases fuel consumption because it reduces the work output
for each cylinder stroke with a given amount of fuel. This happens because the
amount of expansion the combustion gases undergo is reduced since heat is
released later in the power stroke. This also increases exhaust gas temperatures
and reduces power output, but can also lower NOx emissions due to reduced peak
cylinder temperatures. Figures 5.67 through 5.74 compare BSFC and EGT data
obtained from the model baseline and timing retarded model over the entire
operating range of the engine. Figure 5.75 and 5.76 provide maps showing the
BSFC and EGT, respectively, over the operational range of the engine occurring
in the timing retarded model. Both EGT and fuel consumption values are
increased over the entire operating range. Figures 5.77 through 5.80 provide maps
comparing PPMNO (parts per million NO) and BSNO2 (brake specific NO2)
86
emissions from the baseline model and timing retarded model. Note that retarding
the timing 15 degrees greatly reduces NOx output from the engine model.
1000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Retarded Timing
600
500
400
0
5
10
15
20
25
Power (bhp)
Figure 5.67: Timing Retarded 15 deg EGT (K) vs. BHP at 1000 rpm
BSFC (kg/kwhr)
1000 RPM BSFC
0.700
0.600
0.500
0.400
0.300
0.200
0.100
0.000
Wave Baseline
Retarded Timing
0
5
10
15
20
25
Power (bhp)
Figure 5.68: Timing Retarded 15 deg BSFC (kg/kwhr) vs. BHP at 1000 rpm
87
2000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Retarded Timing
600
500
400
0
20
40
60
80
100
Power (bhp)
Figure 5.69: Timing Retarded 15 deg EGT (K) vs. BHP at 2000 rpm
2000 RPM BSFC
BSFC (kg/kwhr)
0.500
0.400
0.300
Wave Baseline
0.200
Retarded Timing
0.100
0.000
0
20
40
60
80
100
Power (bhp)
Figure 5.70: Timing Retarded 15 deg BSFC (kg/kwhr) vs. BHP at 2000 rpm
88
3000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Retarded Timing
600
500
400
0
50
100
150
Power (bhp)
Figure 5.71: Timing Retarded 15 deg EGT (K) vs. BHP at 3000 rpm
BSFC (kg/kwhr)
3000 RPM BSFC
0.700
0.600
0.500
0.400
0.300
0.200
Wave Baseline
Retarded Timing
0.100
0.000
0
50
100
150
Power (bhp)
Figure 5.72: Timing Retarded 15 deg BSFC (kg/kwhr) vs. BHP at 3000 rpm
89
4000 RPM EGT
1000
EGT (K)
900
800
Wave Baseline
700
Retarded Timing
600
500
400
0
50
100
150
200
Power (bhp)
Figure 5.73: Timing Retarded 15 deg EGT (K) vs. BHP at 4000 rpm
4000 RPM BSFC
BSFC (kg/kwhr)
1.200
1.000
0.800
Wave Baseline
0.600
Retarded Timing
0.400
0.200
0.000
0
50
100
150
200
Power (bhp)
Figure 5.74: Timing Retarded 15 deg BSFC (kg/kwhr) vs. BHP at 4000 rpm
90
Figure 5.75: Timing Retarded 15 deg Fuel Consumption Map
Figure 5.76: Timing Retarded 15 deg EGT Map
91
Figure 5.77: Timing Retarded 15 deg PPM NO Map
Figure 5.78: Baseline PPM NO Map
92
Figure 5.79: Timing Retarded 15 deg BSNO2 Map
Figure 5.80: Baseline BSNO2 Map
93
Chapter 6 –Research Summary and Conclusions
Modeling an engine through software is one of the least expensive and quickest
methods of obtaining reasonably accurate data based on reasonably accurate
assumptions. Operating conditions and modifications that would require
significant amounts of time and money to test can be modeled to obtain
information that is accurate enough to make informed decisions and determine
major effects. Information that could not be obtained through conventional
methods can be obtained from a model as well. Modeling the 2.5L Detroit Diesel
engine will help the FutureTruck team to make quick and informed decisions
about which modifications to the engine will help them to best achieve their goals.
The FutureTruck team is very fortunate to have data from an engine dyno to
verify the model against. Without knowing how accurate a model is compared to
the actual engine, the model cannot be used to predict changes in operating
performance with any degree of certainty. Assumptions were made in the
construction of the model because all operating parameters were not obtainable.
However, it can be seen that the model is sufficiently similar to the actual engine
because major characteristics such as brake specific fuel consumption and exhaust
gas temperatures are similar between the model and the dyno.
This research has just scratched the surface in modeling and researching the many
control strategies and modifications available for reduction of emissions while
94
maintaining performance and fuel economy. The benefits and drawbacks of
incorporating each method in a powertrain must be weighed out to develop the
best solution. The main goals are to reduce emissions of NOx, PM, CO, and HC,
while maintaining acceptable fuel economy and power. For each emission one can
take the strategy of minimizing engine out emissions, or maximizing fuel
economy and power and treating the emissions later with aftertreatment.
Generally the former is preferable since it results in less system complexity and
cost. If aftertreatments are used, the system must be designed to provide
appropriate exhaust gas temperatures.
Variable displacement diesel can be ruled out as a major strategy due to its
incredibly high fuel consumption, lack of power, and vibration issues. It really
offers no tangible benefits in terms of emissions.
Thermal throttling only allows for a small increase in exhaust gas temperatures at
low speeds and loads, where the ability to increase the exhaust gas temperatures is
most beneficial if aftertreatments that require high exhaust gas temperatures are
used. The apparatus to perform thermal throttling can also be quite complex. Thus
thermal throttling can be ruled out as a major strategy.
Increasing exhaust backpressure with passive emissions aftertreatment devices,
even though an increase in backpressure is unavoidable, should also not be used
as the main strategy to increase exhaust gas temperatures since backpressure
95
increases with speed and load, and exhaust gas temperatures are not appreciably
raised at low speeds and loads, where it may be beneficial. Increasing exhaust
backpressure also reduces peak power and can dramatically increase fuel
consumption.
Hot EGR provides a thermal throttling effect, as well as increasing peak
combustion temperatures, which can contribute to NOx emissions. However,
cooled EGR has been shown to reduce NOx emissions and improve fuel
consumption, although peak power output is reduced. Since cooled EGR can be
quickly disabled to allow peak power production to be restored and appears to
have no major disadvantages, cooled EGR appears to be a viable emissions
control strategy. The main issue with the use of cooled EGR is ensuring enough
of a pressure differential between the exhaust and intake to maintain sufficient
EGR flow.
Intake throttling has the ability to drastically increase exhaust gas temperatures
when necessary for aftertreatment. This could be very beneficial during cold starts
to warm up the engine and aftertreatment devices more rapidly. Since intake
throttling provides this major benefit and is fairly easy to implement and control,
intake throttling appears to be a viable control strategy.
Retarding the timing can have an extremely beneficial effect on NOx emissions
and can increase exhaust gas temperatures, even though fuel consumption is
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increased. Since timing is easily adjustable and can drastically change fuel
economy and emissions based on speed, load, and amount of EGR, properly
calibrating and perhaps modifying the timing would be an integral part of any
control strategy.
Since sufficient exhaust gas temperatures are needed over a portion of the duty
cycle for diesel particulate filter regeneration, intake throttling could be employed
during periods of idle or low load to raise the EGT. The adverse effects on fuel
economy would be minimized because this would only occur when normal
driving conditions did not put enough load on the engine to raise EGT for a
sufficient amount of time.
The most effective strategy for balancing emissions, fuel economy, performance,
and cost in a diesel engine requires a combination of techniques. Modifications to
the combustion cycle by redesigning the engine and using cooled EGR can reduce
NOx emissions and increase fuel economy. A diesel particulate filter can manage
the remaining emissions problems and act as a muffler. Intake throttling can be
used to selectively raise the exhaust gas temperatures when necessary to
regenerate the diesel particulate filter. This would allow for a robust, low-cost
powertrain solution that effectively controls all emissions while maintaining good
fuel economy.
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Chapter 7 - Future Work
This research leaves open many possibilities for future research based upon
engine modeling.
1) Incorporation of chemical reactions occurring in emissions aftertreatment
devices to obtain tailpipe output emissions information would help
determine aftertreatment effectiveness and viable control strategies.
2) Further improvements in the model incorporating real values for timing,
injection rate, combustion rates, and other parameters would allow more
precise determinations of engine-out emissions such as NOx, HC, CO, and
PM.
3) The engine model could be specifically tuned to offer more power, better
fuel consumption, or reduced emissions. The modifications made to the
engine, such as EGR, could be further optimized.
4) Results from the models could be used to improve the control strategy of
the FutureTruck vehicle in terms of emissions and performance.
Due to this research, the FutureTruck team will now have a reasonably accurate
model of a 2.5L Detroit Diesel engine which can be used for further research.
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4. Fedak, Chris (2003), "Use of WAVE Simulations to Model Performance and
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14. Van Nieuwstadt, Michiel (2003), “Coordinated Control of EGR Valve and
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Appendix
See attached CD for engine model files and data analyses.
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