diesel engine modelling in wave.pdf

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

______________________

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Abstract

The FutureTruck teams 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 Ricardos 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.

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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

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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

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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

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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

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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

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Figure 5.80: Baseline BSNO2 Map 93

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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

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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.

2

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 States strategy incorporates a diesel

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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.

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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

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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

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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

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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 start-

and-stop driving, so modifying the aerodynamics of the vehicle to reduce fuel

consumption would not be very effective.

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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

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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

10

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

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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.

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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

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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.

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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. Chryslers 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 engine-

out 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.

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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.

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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

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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.

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Figure 2.4: Visual representation of hot EGR

Figure 2.5: Visual representation of cooled EGR

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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 engines 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.

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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 teams 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

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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

22

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.

23

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 engine-

out 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.

24

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.

25

Chapter 3 Engine Model

Ricardos 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 one-

dimensional 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.

26

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

27

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.

28

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 masters 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

29

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 manufacturers specifications. Another reason is that test

conditions and methods in the dyno lab may be slightly different than the

manufacturers 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

30

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.

31

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

0

20

40

60

80

100

120

140

160

1000 1500 2000 2500 3000 3500 4000

RPM

Po

wer

(bh

p)

LabWaveRated

Figure 4.1: Peak BHP vs. RPM, Lab vs. Wave vs. Rated

32

BSFC (kg/kwh) - Lab/Wave/Rated

0.000

0.050

0.100

0.150

0.200

0.250

0.300

1000 1500 2000 2500 3000 3500 4000

RPM

BS

FC

(kg

/kw

hr)

LabWaveRated

Figure 4.2: Minimum BSFC vs. RPM, Lab vs. Wave vs. Rated

1000 RPM EGT

400

500

600

700

800

900

1000

0 5 10 15 20 25 30

Power (bhp)

EG

T (K

)

Lab

Wave

Figure 4.3: Baseline model EGT (K) vs. BHP at 1000 rpm

33

1000 RPM BSFC

0.000

0.200

0.400

0.600

0.800

1.000

1.200

0 5 10 15 20 25 30

Power (bhp)

BS

FC (

kg/k

whr

)

Lab

Wave

Figure 4.4: Baseline model BSFC (kg/kwhr) vs. BHP at 1000 rpm

1000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0 5 10 15 20 25 30

Power (bhp)

BS

FC (

kg/k

whr

)

Lab

Wave


34

2000 RPM EGT

400

500

600

700

800

900

1000

0.00 20.00 40.00 60.00 80.00 100.00

Power (bhp)

EG

T (K

)

Lab

Wave


2000 RPM BSFC

0.000

0.200

0.400

0.600

0.800

1.000

0 20 40 60 80 100

Power (bhp)

BS

FC (

kg/k

whr

)

Lab

Wave


35

2000 RPM BSFC

0.000

0.050

0.100

0.150

0.200

0.250

0.300

0 20 40 60 80 100

Power (bhp)

BS

FC (

kg/k

whr

)

Lab

Wave


3000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150

Power (bhp)

EG

T (K

)

Lab

Wave


36

3000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0 50 100 150

Power (bhp)

BS

FC (

kg/k

whr

)

Lab

Wave


4000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150 200

Power (bhp)

EG

T (K

)

Lab

Wave


37

4000 RPM BSFC

0.0000.0500.1000.1500.2000.2500.3000.3500.400

0 50 100 150 200

Power (bhp)

BS

FC (

kg/k

whr

)

Lab

Wave


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.

38

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.

39

Figure 4.13: Baseline Model Fuel Consumption Map

40

Figure 4.14: Baseline Model EGT map

41

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

42

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

400

500

600

700

800

900

1000

0 5 10 15

Power (bhp)

EG

T (K

)

Wave Baseline

Intake Throttle Idle

Figure 5.1: Intake Throttled EGT (K) vs. BHP at 700 rpm

43

700 RPM BSFC

0.0000.1000.2000.3000.4000.5000.6000.700

0 5 10 15

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

Intake Throttle Idle

Figure 5.2: Intake Throttled BSFC (kg/kwhr) vs. BHP at 1000 rpm

1000 RPM EGT

400

500

600

700

800

900

1000

0 5 10 15 20 25

Power (bhp)

EG

T (K

)

Wave Baseline

Intake Throttle


44

1000 RPM BSFC

0.0000.1000.2000.3000.4000.5000.6000.700

0 5 10 15 20 25

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

Intake Throttle


2000 RPM EGT

400

500

600

700

800

900

1000

0 20 40 60 80 100

Power (bhp)

EG

T (K

)

Wave Baseline

Intake Throttle


45

2000 RPM BSFC

0.0000.1000.2000.300

0.4000.5000.6000.700

0 20 40 60 80 100

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

Intake Throttle


3000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150

Power (bhp)

EG

T (K

)

Wave Baseline

Intake Throttle


46

3000 RPM BSFC

0.000

0.200

0.400

0.600

0.800

1.000

0 50 100 150

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

Intake Throttle


4000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150 200

Power (bhp)

EG

T (K

)

Wave Baseline

Intake Throttle


47

4000 RPM BSFC

0.0000.100

0.2000.300

0.4000.500

0.6000.700

0 50 100 150 200

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

Intake Throttle


Figure 5.11: Intake Throttled Fuel Consumption Map

48

Figure 5.12: Intake Throttled EGT Map

49

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.


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

50

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

400

500

600

700

800

900

1000

0 5 10 15 20 25

Power (bhp)

EG

T (K

)

Wave Baseline

Thermal Throttle

Figure 5.13: Thermal Throttled EGT (K) vs. BHP at 1000 rpm

1000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0 5 10 15 20 25

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

Thermal Throttle

Figure 5.14: Thermal Throttled BSFC (kg/kwhr) vs. BHP at 1000 rpm

51

2000 RPM EGT

400

500

600

700

800

900

1000

0 20 40 60 80 100

Power (bhp)

EG

T (K

)

Wave Baseline

Thermal Throttle


2000 RPM BSFC

0.000

0.050

0.100

0.150

0.200

0.250

0.300

0 20 40 60 80 100

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

Thermal Throttle


52

3000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150

Power (bhp)

EG

T (K

)

Wave Baseline

Thermal Throttle


3000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0.600

0 50 100 150

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

Thermal Throttle


53

4000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150 200

Power (bhp)

EG

T (K

)

Wave Baseline

Thermal Throttle


4000 RPM BSFC

0.000

0.200

0.400

0.600

0.800

1.000

0 50 100 150 200

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

Thermal Throttle


54

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.


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

55

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

56

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

400

500

600

700

800

900

1000

0 5 10 15

Power (bhp)

EG

T (K

)

Wave Baseline

VDD Idle

Figure 5.21: Variable Displacement Diesel EGT (K) vs. BHP at 700 rpm

57

700 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0.600

0 5 10 15

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

VDD Idle

Figure 5.22: Variable Displacement Diesel BSFC (kg/kwhr) vs. BHP at 700 rpm

1000 RPM EGT

400

500

600

700

800

900

1000

0 5 10 15 20 25

Power (bhp)

EG

T (K

)

Wave Baseline

VDD


58

1000 RPM BSFC

0.000

0.200

0.400

0.600

0.800

1.000

0 5 10 15 20 25

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

VDD


1300 RPM EGT

400

500

600

700

800

900

1000

0 10 20 30 40

Power (bhp)

EG

T (K

)

Wave Baseline

VDD Idle


59

1300 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0.600

0 10 20 30 40

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

VDD Idle


2000 RPM EGT

400

500

600

700

800

900

1000

0 20 40 60 80 100

Power (bhp)

EG

T (K

)

Wave Baseline

VDD


60

2000 RPM BSFC

0.000

0.200

0.400

0.600

0.800

1.000

0 20 40 60 80 100

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

VDD


3000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150

Power (bhp)

EG

T (K

)

Wave Baseline

VDD


61

3000 RPM BSFC

0.000

0.200

0.400

0.600

0.800

1.000

1.200

0 50 100 150

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

VDD


4000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150 200

Power (bhp)

EG

T (K

)

Wave Baseline

VDD


62

4000 RPM BSFC

0.0000.2000.4000.6000.8001.0001.2001.4001.600

0 50 100 150 200

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

VDD


63

Figure 5.33: Variable Displacement Diesel Fuel Consumption Map

Figure 5.34: Variable Displacement Diesel EGT Map

64

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.

65

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

66

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

400

500

600

700

800

900

1000

0 5 10 15 20 25

Power (bhp)

EG

T (K

)

Wave Baseline

EGR Hot

Figure 5.35: Hot EGR EGT (K) vs. BHP at 1000 rpm

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1000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0.600

0 5 10 15 20 25

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

EGR Hot

Figure 5.36: Hot EGR BSFC (kg/kwhr) vs. BHP at 1000 rpm

2000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150

Power (bhp)

EG

T (K

)

Wave Baseline

EGR Hot


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2000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0.600

0 50 100 150

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

EGR Hot


3000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150

Power (bhp)

EG

T (K

)

Wave Baseline

EGR Hot


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3000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0.600

0 50 100 150

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

EGR Hot


4000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150 200

Power (bhp)

EG

T (K

)

Wave Baseline

EGR Hot


70

4000 RPM BSFC

0.0000.1000.2000.3000.4000.5000.6000.7000.800

0 50 100 150 200

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

EGR Hot


Figure 5.43: Hot EGR % of Exhaust Gas in Intake Map

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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

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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

400

500

600

700

800

900

1000

0 5 10 15 20 25

Power (bhp)

EG

T (K

)

Wave Baseline

EGR Cooled

Figure 5.44: Cooled EGR EGT (K) vs. BHP at 1000 rpm

1000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0.600

0 5 10 15 20 25

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

EGR Cooled

Figure 5.45: Cooled EGR BSFC (kg/kwhr) vs. BHP at 1000 rpm

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2000 RPM EGT

400

500

600

700

800

900

1000

0 20 40 60 80 100

Power (bhp)

EG

T (K

)

Wave Baseline

EGR Cooled


2000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0.600

0 20 40 60 80 100

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

EGR Cooled


74

3000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150

Power (bhp)

EG

T (K

)

Wave Baseline

EGR Cooled


3000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0.600

0 50 100 150

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

EGR Cooled


75

4000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150 200

Power (bhp)

EG

T (K

)

Wave Baseline

EGR Cooled


4000 RPM BSFC

0.0000.1000.2000.3000.4000.5000.6000.7000.800

0 50 100 150 200

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline

EGR Cooled


76

Figure 5.52: Cooled EGR % of Exhaust Gas in Intake Map

Figure 5.53: Cooled EGR PPM NO Map

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Figure 5.54: Baseline PPM NO Map

Figure 5.55: Cooled EGR BSNO2 Map

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Figure 5.56: Baseline BSNO2 Map

79

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

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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.

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1000 RPM EGT

400

500

600

700

800

900

1000

0 5 10 15 20 25

Power (bhp)

EG

T (K

) Wave Baseline

700 mbarbackpressure

Figure 5.57: 700 mbar max backpressure EGT (K) vs. BHP at 1000 rpm

1000 RPM BSFC

0.0000.1000.2000.3000.4000.5000.6000.700

0 5 10 15 20 25

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline


Figure 5.58: 700 mbar max backpressure BSFC (kg/kwhr) vs. BHP at 1000 rpm

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2000 RPM EGT

400

500

600

700

800

900

1000

0 20 40 60 80 100

Power (bhp)

EG

T (K

) Wave Baseline



2000 RPM BSFC

0.0000.050

0.1000.1500.2000.250

0.3000.350

0 20 40 60 80 100

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline



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3000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150

Power (bhp)

EG

T (K

) Wave Baseline



3000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0 50 100 150

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Baseline



84

4000 RPM EGT

400

500

600

700

800

900

1000

0 50 100 150 200

Power (bhp)

EG

T (K

) Wave Baseline



4000 RPM BSFC

0.000

0.100

0.200

0.300

0.400

0.500

0 50 100 150 200

Power (bhp)

BS

FC (

kg/k

whr

)

Wave Std



85

Figure 5.65: 700 mbar max Backpressure Exhaust System Pressure Map

Figure 5.66: 700 mbar max Backpressure EGT Map

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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)

87

emissions from the baseline model and timing retarded model. Note that retarding

the timing

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