STEADY STATE AND TRANSIENT EFFICIENCIES OF A

FOUR CYLINDER DIRECT INJECTION DIESEL ENGINE

FOR IMPLEMENTATION IN A HYBRID ELECTRIC VEHICLE

A Thesis

Presented to

The Graduate Faculty of The University of Akron

In Partial Fulfillment

of the Requirements for the Degree

Masters of Science

Charles Van Horn

August, 2006 STEADY STATE AND TRANSIENT EFFICIENCIES OF A

FOUR CYLINDER DIRECT INJECTION DIESEL ENGINE

FOR IMPLEMENTATION IN A HYBRID ELECTRIC VEHICLE

Charles Van Horn

Thesis

Approved: Accepted:

Advisor Department Chair Dr. Scott Sawyer Dr. Celal Batur

Faculty Reader Dean of the College Dr. Richard Gross Dr. George K. Haritos

Faculty Reader Dean of the Graduate School Dr. Iqbal Husain Dr. George R. Newkome

Date

ii ABSTRACT

The efficiencies of a four cylinder direct injection diesel engine have been

investigated for the implementation in a hybrid electric vehicle (HEV). The engine was cycled

through various operating points depending on the power and torque requirements for the HEV.

The selected engine for the HEV is a 2005 Volkswagen 1.9L diesel engine. The 2005 Volkswagen

1.9L diesel engine was tested to develop the steady-state engine efficiencies and to evaluate the

transient effects on these efficiencies. The peak torque and power curves were developed using a

water brake dynamometer. Once these curves were obtained steady-state testing at various

engine speeds and powers was conducted to determine engine efficiencies. Transient operation of

the engine was also explored using partial throttle and variable throttle testing. The transient

efficiency was compared to the steady-state efficiencies and showed a decrease from the steady-

state values. Changes in engine efficiency and how it impacts vehicle fuel economy for steady speeds was also investigated.

From the steady-state and transient testing suggested operating points for the engine implementation in the series-parallel HEV developed by The University of Akron were made. The steady-state efficiency data is useful for the determination of operation points for series, parallel, and spilt hybrid modes. Transient efficiencies behavior is useful during acceleration of the vehicle at both high and low speeds, as well as the transition between hybrid operating modes. Engine operating points for other applications may also be derived from this data.

iii ACKNOWLEDGEMENTS

I would like to thank the following people for their contributions to this project:

- Dr. Sawyer, of the Mechanical Engineering Department at the University of Akron, for his

guidance and help through this project as well as my graduate and undergraduate studies.

- Dr. Gross, of the Mechanical Engineering Department at the University of Akron, for his

guidance and development throughout this project.

- Dr. Husain, of the Electrical Engineering Department at the University of Akron, for his

leadership of the Challenge X program as well as his advisement on this paper.

- Mr. Steve Gerbetz for his guidance and assistance with the dynamometer test setup and

many other technical issues.

- The entire University of Akron Challenge X Team, students, faculty, and staff for making

this project a success and allowing me the opportunity to do this research.

iv TABLE OF CONTENTS

Page

LIST OF TABLES vii

LIST OF FIGURES viii

CHAPTER

I. INTRODUCTION 1

1.1 The University of Akron’s Hybrid Electric Vehicle 2

1.2 The University of Akron’s Hybrid Electric Vehicle Engine Selection 4

1.3 Engine Use in a Hybrid Electric Vehicle 9

1.4 Thesis Overview 9

II. BACKGROUND OF THE STUDY 11

2.1 Historical Survey of Previous Work in the Area 11

III. PEAK ENGINE POWER AND TORQUE CURVES 15

3.1 Test Setup 15

3.2 Acceleration Testing 18

IV. STEADY-STATE ENGINE EFFICIENCY 25

4.1 Test Setup 25

4.2 Steady-State Efficiency Calculations 28

4.3 Efficiency Uncertainty 40

4.4 Engine Operating Point 41

v V. TRANSIENT ENGINE EFFICENCY 44

5.1 Partial Throttle Testing 45

5.2 Varying Throttle Testing 50

VI. STEADY-STATE AND TRANSIENT ENGINE EFFICENCY COMPARISION 59

6.1 Steady-State and Partial Throttle Comparisons 59

6.2 Steady-State and Varied Throttle Comparisons 63

6.3 Steady-State and Transient Fuel Economy Comparison 66

VII. SUMMARY 68

REFERENCES 70

APPENDICIES 72

APPENDIX A. VOLKSWAGEN ENGINE DATA 73

APPENDIX B. EXPERIMENTAL TEST SETUP 74

APPENDIX C. CONTACT INFORMATION 77

vi LIST OF TABLES

Table Page

4.1.1 Engine Test Speeds and Power 26

4.1.2 Testing Loads and Throttle Percentages for 1500 RPM 28

4.2.1 Steady-State Efficiency Repeatability for 10 HP at 1500 RPM 37

4.2.2 Steady-State Efficiency Repeatability for 20 HP at 2000 RPM 37

4.2.3 Steady-State Efficiencies 39

5.1.1 Engine Efficiency and Uncertainty at 30% of Full Throttle 46

5.1.2 Engine Efficiency and Uncertainty at 40% of Full Throttle 47

5.1.3 40% Partial Throttle Efficiency Repeatability 49

vii LIST OF FIGURES

Figure Page

1.1 The University of Akron Hybrid Electric Vehicle Architecture 4

1.2 Internal Combustion Engine Piston-Cylinder 5

1.3 Pressure-Volume Diagram for Otto Cycle 6

1.4 Pressure-Volume Diagram for Ideal Diesel Cycle 7

3.1.1 Superflow SF-901 Dynamometer Engine Test Stand 16

3.1.2 Superflow SF-901 Dynamometer Console 17

3.1.3 VW 1.9L Engine Coupled to SF-901 Dynamometer 18

3.2.1 Power and Torque Curves for Test 1 20

3.2.2 STP Corrected Power and Torque Curves for Test 1 20

3.2.3 Targeted Engine Speed vs. Actual Speed for Test 1 21

3.2.4 Engine Power and Torque Curve Comparisons for the Four Tests 22

3.2.5 Total Averaged Engine Power and Torque Curves 23

3.2.6 STP Corrected Total Average Power and Torque Curves 24

4.1.1 Fuel Consumption for Engine Power of 10 HP at 1500 RPM

and 50 HP at 2000 RPM 27

4.1.2 Temperature-Entropy Diagram for Carnot and Ideal Diesel Cycles 30

4.2.1 Steady-State Efficiency Points plotted for Engine Power 32

4.2.2 Steady-State Efficiency Points plotted for Engine Torque 32

viii 4.2.3 Best Efficiency Curve 33

4.2.4 Engine Efficiency for 10 HP at Various Engine Speeds 34

4.2.5 Engine Efficiency for 20 HP at Various Engine Speeds 34

4.2.6 Engine Efficiency for 30 HP at Various Engine Speeds 35

4.2.7 Engine Efficiency for 40 HP at Various Engine Speeds 35

4.2.8 Engine Efficiency for 50 HP at Various Engine Speeds 36

4.2.9 Engine Efficiency for 60 HP at Various Engine Speeds 36

4.3.1 Recorded Engine Power for 1500 RPM 40

4.4.1 Engine Operating Temperature 42

4.4.2 Engine Operating Temperature without Thermal Runaway and Overcooling 42

5.1.1 Engine Power and Efficiency at 30% of Full Throttle 48

5.1.2 Engine Power and Efficiency at 40% of Full Throttle 48

5.2.1 Throttle Voltage Input for Transient Testing at 1500 RPM 50

5.2.2 Throttle Voltage Input for Transient Testing at 1750 RPM 51

5.2.3 Torque Relationship for 2000 RPM 52

5.2.4 Transient Efficiencies for 1500 RPM 54

5.2.5 Transient Efficiencies for 1750 RPM 55

5.2.6 Transient Efficiencies for 2000 RPM 56

5.2.7 Transient Efficiencies for 2500 RPM 57

6.1.1 Steady-State and 30% Partial Throttle Comparison for 1500 RPM 60

6.1.2 Steady-State and 30% Partial Throttle Comparison for 2000 RPM 60

6.1.3 Steady-State and 30% Partial Throttle Comparison for 2500 RPM 61

6.1.4 Steady-State and 40% Partial Throttle Comparison for 1500 RPM 61

ix 6.1.5 Steady-State and 40% Partial Throttle Comparison for 2000 RPM 62

6.1.6 Steady-State and 40% Partial Throttle Comparison for 2500 RPM 62

6.2.1 Steady-State and Varied Throttle Efficiency Comparison for 1500 RPM 64

6.2.2 Steady-State and Varied Throttle Efficiency Comparison for 1750 RPM 64

6.2.3 Steady-State and Varied Throttle Efficiency Comparison for 2000 RPM 65

6.2.4 Steady-State and Varied Throttle Efficiency Comparison for 2500 RPM 65

B.1 Modified Transmission Bell Housing 74

B.2 Modified Flywheel 75

B.3 Custom Bell Housing Extension 75

B.4 Modified Innovations Engineering Torque Dampener 76

B.5 Innovations Engineering Dynamometer Adapter 76

x CHAPTER I

INTRODUCTION

The University of Akron was selected in 2004 to participate in Challenge X, a North

American vehicle design competition. The Department of Energy and General Motors Company are the headline sponsors of Challenge X. The goal of the competition is to re-engineer a 2005

Chevrolet Equinox to minimize energy consumption and emissions while maintaining the utility of the vehicle. During the first part of the 2004 school year the team developed a hybrid electric vehicle (HEV) architecture for use in the competition. This development process included a literature review into current and future HEV designs, vehicle simulations using the Powertrain

Simulation Analysis Toolkit (PSAT), Greenhouse Gasses Regulated Emissions and Energy Usage

Toolkit (GREET), as well as hand calculated derivations. The team also used several marketing research surveys conducted by marketing research groups from The University of Akron’s College of Business Administration.

This thesis presents the results of experiments conducted and analysis for the characterization of the internal combustion engine selected for use in The University of Akron’s

HEV architecture. The engine was cycled through various operating points depending on the power and torque requirements for the HEV. The selected engine, a 2005 Volkswagen 1.9L diesel engine, was tested to develop the steady-state engine efficiencies and to evaluate the transient effects on these efficiencies. The peak torque and power curves were developed using a water

1 brake dynamometer. Once these curves were obtained steady-state testing at various engine

speeds and powers was conducted to determine engine efficiencies. Transient operation of the engine was also explored using partial throttle and variable throttle testing. The transient efficiency was compared to the steady-state efficiencies and showed a decrease from the steady-state values. Changes in engine efficiency and how it impacts vehicle fuel economy for steady speeds was also investigated.

From the steady-state and transient testing suggested operating points for the engine implementation in the series-parallel HEV developed by The University of Akron were made. The steady-state efficiency data is useful for the determination of operation points for series, parallel, and spilt hybrid modes. Transient efficiencies behavior is useful during acceleration of the vehicle at both high and low speeds, as well as the transition between hybrid operating modes. Engine operating points for other applications may also be derived from this data.

1.1 The University of Akron’s Hybrid Electric Vehicle

There are several different conventional hybrid vehicle architectures currently in use.

Parallel hybrid architectures use one electric machine, which provides assistance to the Internal

Combustion Engine (ICE). The electric machine takes power from the battery to provide assistance to the ICE and once the battery state of charge has been depleted, the electric machine is used in a regenerative mode to charge the battery. In a parallel architecture the operating range

2 of the ICE is varied because it operates through a range similar to a conventional vehicle with assistance from the electric machine. The Honda Insight is an example of production HEVs with a parallel architecture.

Series hybrid architectures use an ICE coupled to a generator, which supplies a drive motor. The ICE can be used in its most efficient operating range. All of the series powertrain components must be sized for the maximum needed power. Since all components in a series vehicle must be sized to the same power they are inherently heavy and there are currently no passenger vehicles using this technology. However, GM has employed a series architecture into public transportation buses. Parallel architectures tend to exhibit very good performance, while a series architecture exhibits better fuel economy and reduced emissions.

More advanced HEVs combine the two technologies of series and parallel. The Toyota

Prius is an example of a split architecture, which combines series and parallel technology. It is more complicated and requires an additional mechanical linkage as compared to a series architecture and has one additional electric machine as compared to parallel architecture.

From this knowledge of current HEVs and additional research and modeling, the University of Akron team designed a combination of a parallel and a series architecture, referred to as a series-parallel (Fig. 1.1). This architecture consists of an ICE driving the front wheels mechanically and the rear wheels electrically through a generator and a drive motor. This architecture has the performance of the parallel architecture as well as the fuel economy of the series architecture. The series-parallel architecture allows for great flexibility in the HEV modes of operation. The vehicle can be operated in an electric only mode, a mechanical only mode, a series mode, a parallel mode, and a split mode. In the electric only mode energy from the energy storage system would be used to drive the rear motor. The ICE engine and coupled transmission could be used to drive the front

3 wheels with the electrical system not operating and give a mechanical only mode. The series

mode would utilize the ICE running at an optimal point and sending power through the generator to

the rear drive motor. The drive motor can also provide assist to the engine during an acceleration

event, which would be a parallel mode. Finally the engine could be used to both power the front

wheels and also send power through the generator to the rear drive motor in a split operation

mode.

Figure 1.1. The University of Akron Hybrid Electric Vehicle Architecture.

1.2 The University of Akron’s Hybrid Electric Vehicle Engine Selection

After the architecture was developed the individual components for the system had to be selected. The type and size of the ICE, which is the main power source for the vehicle, had to be

4 determined. Internal combustion engines are broken up into two different categories, spark ignition

(SI) and compression ignition (CI). For the two different types of ICEs the main components of the engines are essentially the same.

Figure 1.2. Internal Combustion Engine Piston-Cylinder.

The ICE consists of a piston in a cylinder that is attached to crankshaft (Figure 1.2). The combustion process occurs in the cylinder and pushes the piston downwards. The piston then transfers this power through a connecting rod to the crankshaft which provides the rotational torque to drive the vehicle. In an SI engine fuel and air are mixed in the cylinder and ignited by a spark plug. In a CI engine air is compressed to a high enough temperature and pressure so that when the fuel is injected it combusts spontaneously. Both ICEs considered were four stroke engines.

The four strokes are the intake, compression, power, and exhaust strokes.

To better understand the differences between a SI engine and a CI engine the ideal cycles can be compared. The Otto cycle is an ideal cycle that models an SI engine consisting of four

5 internally reversible processes. Figure 1.3 illustrates the four processes in a pressure-volume

diagram.

Figure 1.3. Pressure-Volume Diagram for Otto Cycle.

From points 1 to 2 is an isentropic (constant entropy, or without heat transfer), compression of the air as the piston moves from bottom dead center (BDC) to top dead center (TDC). A constant volume heat transfer to the air from an external source while the piston is at TDC is represented from points 2 to 3. This process represents the ignition of the air fuel mixture and the rapid burning. From points 3 to 4 is an isentropic expansion from TDC to BDC or the power stroke.

Finally from points 4 to 1 is a constant volume heat rejection from the air while the piston is at BDC.

The thermal efficiency of the Otto cycle is only a function of the compression ratio. As the compression ratio increases the thermal efficiency increases. With an SI engine as the compression ratio increases so does the likelihood of autoignition or “knock”. When a portion of the air fuel mixture is ignited the remaining mixture experiences an increase in pressure and temperature. If this temperature becomes high enough then the air fuel mixture can autoignite before it is consumed by the flame front. By increasing the compression ratio the mixture

6 temperature in the compression stroke increases and this leads to a higher chance of autoignition.

This limits the thermal efficiency of an SI engine.

Figure 1.4. Pressure-Volume Diagram for Ideal Diesel Cycle.

The ideal Diesel cycle differs from the Otto cycle and can be seen in Figure 1.4. From

points 1 to 2 is still an isentropic compression. Points 2 to 3 represent a constant pressure heat

addition which is the beginning of the power stroke. Points 3 to 4 represent an isentropic expansion and the remainder of the power stroke. Finally points 4 to1 represent a constant volume heat rejection at BDC similar to the Otto cycle. The thermal efficiency of a CI engine is also a

function of the compression ratio. Although the efficiency of the Diesel Cycle will be less than the

efficiency of a Otto Cycle at the same compression ratio, diesel engines can be operated at higher

compression ratios without concern for autoignition. This higher compression ratio allows for

higher efficiencies for a CI engine.

From these ideal cycles it can be seen that a CI engine can obtain higher efficiencies and

would be favorable for use in an application where fuel economy is a concern. With this knowledge

as well as energy calculations and the University’s experience with the available competition fuels,

7 a CI engine was selected for use in Akron’s architecture. The rules of the Challenge X competition allow for the usage of four fuels: B20 biodiesel, E85 ethanol, reformulated gasoline, and hydrogen.

The University did not feel confident in the usage of hydrogen. The remaining three fuels were

then compared on a basis of fuel economy with emissions considerations. B20 biodiesel was

selected based on the high energy content of B20 biodiesel coupled with the high efficiency of a

diesel engine.

A study into several diesel engines available in the United States as well as Europe was

conducted by the Akron team. Sizing of the diesel engine was based on the towing competition

event. The vehicle must be able to tow a 2500 lb trailer up a six percent grade at 55 mph while

sustaining charge. Based on the charge-sustaining requirement the Akron team decided to pursue

an engine adequately sized to perform the trailer tow event without assistance from the electrical

drive system. The power requirement of 108 HP would be required for the towing event. With all

these design parameters a 2005 Volkswagen (VW) 1.9L Turbocharged Direct Injection engine was

selected for use in the series-parallel architecture. Specifications for the 2005 VW 1.9L engine can

be found in Appendix A.

The 2005 VW engine rated at 100 HP has a unit injection system, which differs from the

2004 VW engine. An overhead camshaft drives four unit injectors that pressurize the fuel and

inject it directly into the cylinders. The 2005 VW engine also uses a Garret variable vane

turbocharger, which is another improvement over previous generations of this engine.

In the architecture, the engine will be operated in series at low speeds and the engine will

also be used as the primary power source for highway speeds. ICEs are inherently efficient at

highway speeds. In order to optimize the fuel economy of the vehicle during city, or low speeds,

the engine will need to operate at points of highest efficiency. Since the architecture has many

8 different operating modes the engine may be accelerating of decelerating out of a mode and to a

point of high efficiency. The study of how the engine’s efficiency changes due to this acceleration or deceleration is also important to overall fuel economy.

This engine is relatively new to the United States and the data on the engine is limited.

This study is intended to develop more information about this engine that can be used in multiple applications including use in an HEV.

1.3 Engine Use in a Hybrid Electric Vehicle Application

The efficiency testing can be used to determine the best operation points for the diesel engine depending on the operation mode. The steady-state efficiency data would be useful for operation in series, parallel, and spilt hybrid modes. Transient efficiencies would be useful during acceleration of the vehicle at both high and low speeds, as well as the transition between hybrid operating modes. The vehicle supervisory controller’s task is to use these efficiencies developed in this thesis, as well as the characteristics of the electrical machines, and the emissions characteristics of the engine to find optimized vehicle operation points.

1.4 Thesis Overview

This thesis shows the development of the efficiency points of the 1.9L VW engine for use in the series-parallel HEV architecture. Chapter II discusses the previous, current, and future research in the area of diesel engines. In Chapter III the test set-up is introduced and the maximum torque and power curves are developed using standard diesel fuel to verify published data. The power and torque curves are also used to develop testing points for determining the steady-state efficiency. Steady-state testing and the determination of efficiency points for the

9 engine are discussed in Chapter IV. This chapter also illustrates how the efficiency points can be

used to determine operating points for the University of Akron’s HEV architecture. A relationship

between the engine efficiency and vehicle fuel economy at a steady speed is developed in Chapter

IV. Transient efficiencies are explored in Chapter V by using partial and varied throttle testing. A comparison between the transient and steady-state efficiencies is discussed in Chapter VI. Finally,

Chapter VII summarizes the results of the study and suggests future work.

10 CHAPTER II

BACKGROUND OF THE STUDY

2.1 Historical Survey of Previous Work in the Area

The compression ignition engine was first developed by Rudolf Diesel in 1892 [1]. Rudolf

Diesel was looking to create an engine that was more efficient than the spark ignition engines

during that time. A diesel engine intakes air into the cylinder and it is compressed by a rising

piston. As the air is compressed, its temperature increases. Once the piston has reached the top

of its stroke fuel is injected into the combustion chamber. The fuel is injected at a high pressure

through an atomizing nozzle to ensure a more complete combustion. The fuel and air mixture then

ignites due to the high temperature of the air, and rapidly burns. This combustion causes the

gases in the combustion chamber to rapidly heat and expands pushing the piston down. The

connecting rod transmits this motion to the crankshaft which then provides rotational power for the

engine. The exhaust gas is then pushed out of the cylinder and a fresh charge of air is taken in

through either valves or ports. Diesel engines lend themselves to forced induction. In most diesel applications a turbocharger or supercharger is used. An intercooler may also be used to increase the power output of the engine by increasing the amount of available oxygen for the combustion

process. Diesel engines have compression ratios that range from 14:1 to as high as 25:1 [1]. Most

diesel engines also use direct fuel injection which means that the fuel is injected

11 directly into the cylinder. Some diesel engines also have glow plugs to help the combustion

process in cold temperatures. A glow plug is an electrically heated wire that helps ignite the fuel when the engine temperature is not high enough to ignite the fuel.

Diesel engines have many different areas of use. Fixed installation engines, cars and light commercial vehicles, heavy goods vehicles, construction and agricultural machinery are just a few examples of diesel engine applications [1]. For cars fast running diesel engines are used and their

sizes can range from 3 cylinders to 10 cylinders with a variety of displacements. Ford, General

Motors, DaimlerChrylser, and Volkswagen offer diesel engines in some of their vehicles [2].

Most vehicle diesel design efforts are targeted towards having a maximum torque at low

engine speeds [1]. This allows for the most economical fuel consumption as well as a preferable

engine response. Typical diesel engines have high torque at lower speeds and their torque curve

is flatter than that of an SI engine. There are many advantages in using a CI engine over an SI

engine. As previously discussed diesel engines are more efficient than SI engines due to the

higher compression ratio [3]. Diesel fuel also has a higher energy content than standard gasoline.

Diesel engines have to be built stronger than SI engines because of the high combustion

pressures. The lack of an electrical ignition system increases the reliability of a diesel engine.

There is both an advantage and disadvantage because diesel engines tend to last much longer

than SI engines but they are also much heavier. Since diesel engines are not a high revolution

engine, greater displacements are needed to provide the same power as an SI engine.

There have been many tests of diesel engines and their incorporation into future vehicles

and HEVs [4, 5, 6, 7]. In most HEV applications, there is a tradeoff between fuel economy and

reduced emissions. Corporate Average Fuel Economy (CAFE) standards, worries about global

warming, and decreasing the dependency on foreign oil are all factors pushing for high fuel

12 economy [8]. Advantages of a compression ignition engine versus a spark ignition engine are a higher compression ratio, greater excess of air, and no throttle-related losses in the part-load range

[1]. The use of a CI engine in an HEV provides additional optimization potential. Ford,

DaimlerChrysler, and GM all have prototype diesel hybrid vehicles as a part of the Partnership for the Next Generation Vehicle [4]. All three prototypes show a mile per gallon percentage gain of over 27% just by dieselization [4].

Previous design iterations of the 1.9L VW engine, engine code ALH, have been used in many tests. The ALH engine has a different fuel delivery system using a high pressure pump connected to the crankshaft with four high pressure fuel lines from the pump to four injectors in each cylinder. The 2005 1.9 VW engine, engine code BEW, has unit injectors for each cylinder.

The unit injector system can achieve the highest injection pressure of all diesel injection systems in use today [1]. Diesel engines using unit injection systems tend to produce low emissions, high fuel economy, good performance, and run quieter than other injection systems. They are actuated by an overhead camshaft. A very high fuel injection pressure can be achieved because there are no high pressure fuel lines [1]. Advances in this unit injection system are still being made.

Refinements of the electronic control, as well as the development of higher pressure injection are two areas of expansion for the unit injection system [1]. The new engine with the unit injectors has only been available in the US for two years now so the amount of published test data for this engine is very limited. Argonne National Laboratory has developed a software package,

Powertrain System Analysis Tool Kit (PSAT), for the development of HEV architectures and testing. Steady-State test data from Oakridge National Laboratories as well as Argonne National

Laboratories on previous versions of the 1.9L is available within the software for engine modeling purposes.

13 Research into the effects of transient load conditions has been conducted on indirect injection (IDI) diesel engines [9]. A model of a six-cylinder IDI was created and the engine’s response to a step input was modeled and compared to a test engine. This study found that turbocharger inertia, fuel pump operation, and exhaust manifold volume can have significant effects on the engines transient response.

The current research trend in diesel applications focuses on emissions. Due to the inherent high efficiency of a diesel engine most research has looked into the emissions issues that are related to a diesel engine. Future development of diesel engines includes the further development of variable vane turbocharging systems and high pressure fuel injection systems.

Other research areas that are expanding involve the use of alternative fuels in diesel engines

[10, 11, and 12]. Biodiesel and compressed natural gas, as well as dual fuel usages are some examples of alternative fuels being used in diesel engines.

14 CHAPTER III

PEAK ENGINE POWER AND TORQUE CURVES

Baseline maximum torque and horsepower curves for the 1.9L VW engine were needed to first verify published power and torque numbers. This testing was also used to determine test points for the steady-state data based on the maximum power curve. The test setup is described in detail and the torque and power curves are developed using four separate acceleration tests.

Finally the uncertainty in the collected and calculated data is discussed.

3.1 Test Setup

A test set up using a water brake dynamometer was used to develop the curves

(Appendix B). A SuperFlow SF-901 water brake dynamometer with WinDyn software was used to develop the power and torque curves (Fig. 3.1.1). The fluid brake dynamometer is an agitator type where baffles are used to change the momentum of the fluid to produce the braking force. The water brake, or power absorber, provides the ability to handle torques of up to 1000 LB-FT and can operate up to 10,000 RPM with a maximum dissipation of 1000 HP. The dynamometer always had a fresh supply of building water, which helps keep the brake at a consistent temperature. The dynamometer console can display and record data from the various sensors on the engine test stand (Fig. 3.1.2). There are multiple gages and meters to allow for the monitoring of

15 temperatures, voltages, horsepower, torque, engine speed, and fluid flow rates. The console has a

Motorola 6809 microprocessor for control and data acquisition. The memory of the console is

limited to 72 recordings. The engine speed is measured using a magnetic pick up on a 60-tooth gear. The engine torque is measured using a strain gauge bridge that can read to 1000 LB-FT.

The console then calculates the power.

New torque and speed sensors were installed and the dynamometer was calibrated to

300 LB-FT before testing began. The capacity valve on the dynamometer was set to allow for the

proper range of speed and torque for the VW 1.9L engine.

Figure 3.1.1. Superflow SF-901 Dynamometer Engine Test Stand.

16

Figure 3.1.2. Superflow SF-901 Dynamometer Console.

The engine, from a 2005 Volkswagen Beetle, had 304 miles of in-vehicle testing before it was taken for dynamometer testing. The complete vehicle wiring harness was removed from the vehicle along with the engine and transmission. The transmission was then removed from the engine so that it could be directly coupled to the dynamometer through a drive shaft. A modified transmission bell housing was extended and bolted using two ½ inch plates and a 6 in diameter tube to the dynamometer engine support plate (Appendix B). The engine support plate is attached to the engine stand through a shock mount. A torque dampening plate and drive shaft purchased from Innovations Engineering were used for the coupling (Appendix B). By extending the bell housing the engine torque is reacted back through the bell housing to the engine instead of through

the engine stand. An initial attempt at reacting the torque through the engine stand provided less

than favorable results so the bell housing extension was machined. The front of the engine was

supported using a custom frame and the stock Volkswagen motor mount. Some vibration

absorption material was used between the custom frame and the engine stand.

17

Figure 3.1.3. VW 1.9L Engine Coupled to SF-901 Dynamometer.

The engine setup was kept as close to stock as possible. The wiring harness was striped of any extraneous items such as headlights and interior electronics. The stock electronic throttle control, air cleaner, intake, radiator, and exhaust components were used for the testing. The purpose was to keep the engine setup close to what would be used in the final vehicle. Figure 3.1.3 shows the engine and dynamometer test setup.

3.2 Acceleration Testing

In order to obtain peak torque and power measurements for the engine an acceleration test was conducted. The throttle was incremented to full throttle as the brake put a load on the engine. The starting engine speed was 1200 RPM. This avoided the rough idle of the engine and allowed for more accurate measurements. The stop engine speed was set at 4200 RPM. The 18 stop speed was set below the engine red line to avoid damage to the engine. The dynamometer

was operated in the servo load control mode, which decreased the load so that the correct engine

acceleration of 100-RPM per second could be achieved. At each step a measurement of torque

was recorded and from that horsepower could be calculated. Once the upper test speed was

reached the dynamometer automatically applied full load to return the engine to the initial starting

speed. The intake air temperature, barometric pressure, and vapor pressure of the test cell were also measured to calculate a Standard Temperature and Pressure (STP) correction factor. The

STP used are 60 F, and 29.92 in Hg, dry air [13]. The STP correction factor is calculated as:

()459.7 + Air Temp  29.92  (3.2.1) STP Correction Factor = ×   536.7  Barometric Pressure - Vapor Pressure 

where the measured air temperature, barometric pressure and vapor pressure of the test cell are input into the equation. The power and torque are then multiplied by this factor. Figure 3.2.1 shows the uncorrected torque and power found during the first acceleration test. Figure 3.2.2 shows the STP corrected torque and power found during the first acceleration test. Figure 3.2.3 shows the target speed versus the actual engine speed. A line with a slope of one for this figure shows that the dynamometer is correctly calibrated and that the recorded data is accurate [13].

19 100 200 90 180 80 160 70 140 60 120 50 100

Power (HP) Power 40 80 Torque (LB-FT) 30 60 20 40 10 20 0 0 0 500 1000 1500 2000 2500 3000 3500 4000 4500 Engine Speed (RPM)

Measured Power Measured Torque

Figure 3.2.1. Power and Torque Curves for Test 1.

120 200 180 100 160 140 80 120 60 100 80 Power (HP) Power 40 60 (LB-FT) Torque

20 40 20 0 0 0 500 1000 1500 2000 2500 3000 3500 4000 4500 Engine Speed (RPM)

STP Corrected Power STP Corrected Torque

Figure 3.2.2. STP Corrected Power and Torque Curves for Test 1.

20 4500

4000

3500

3000

2500

2000

1500

Actual Engine Speed (RPM) 1000

500

0 0 500 1000 1500 2000 2500 3000 3500 4000 4500 Targeted Engine Speed (RPM)

Figure 3.2.3. Targeted Engine Speed vs. Actual Speed for Test 1.

The acceleration test was run four times to ensure accuracy and repeatability of the data.

Each test was analyzed and compared and an average of the four tests was used to create the baseline maximum torque and power curves. Figure 3.2.4 shows the four test results overlapped.

The relative uncertainty was calculated using the following equation:

   1 dy  ei = ×U x (3.2.1)   i y dxi  xi =xi  where y is the governing equation, x is the variable with some uncertainty, and U is half of the i xi

resolution of the measurement device. The total relative uncertainty ey is calculated as:

2 2 2 2 e y = e1 + e2 + e3 + ... + ei (3.2.2)

21 The maximum relative uncertainty for the power and torque for each of the four tests was 0.07%.

This uncertainty is so low because the resolution of the measured torque is 0.1 LB-FT and the resolution of the measured speed is 1 RPM. The four tests were compared at each speed and the standard deviation was calculated. The largest difference in values occurs at 2000 RPM. Between

these four tests the standard deviation is 3.1 HP. To try to estimate an error for the averaged

values the standard deviation was divided by the mean and multiplied by 100 to give 4.8%. A

standard deviation of 2.7 occurred at 3600 RPM. All the other data points had a much lower

standard deviation.

200 200 180 180 160 160 140 140 120 120 100 100 80 80 Power (HP) Power 60 60 Torque (LB-FT) 40 40 20 20 0 0 0 1000 2000 3000 4000 5000 Engine Speed (RPM)

Figure 3.2.4. Engine Power and Torque Curve Comparisons for the Four Tests.

The four tests were then averaged for both the uncorrected and STP corrected data. Figure 3.2.5

shows a maximum power is 94.7 HP at 4000 RPM and the maximum torque is 178.1 LB-FT at

2300 RPM. Figure 3.2.6 shows the STP corrected values are 97.5 HP at 4000 RPM 181.6 LB-FT

22 at 2300 RPM. The Volkswagen published numbers are 100 HP at 4000 RPM and 180 LB-FT at

1800-2400 RPM [14]. Both the STP corrected and uncorrected values for horsepower and torque are very close to the published data.

200 100 180 90 160 80 140 70 120 60 100 50 80 40 Power (HP) 60 30 (LB-FT) Torque 40 20 20 10 0 0 0 1000 2000 3000 4000 5000 Engine Speed (RPM)

Measured Torque Measured Power

Figure 3.2.5. Total Averaged Engine Power and Torque Curves.

23 120 200 180 100 160 80 140 120 60 100 80 Power (HP) Power 40 60 Torque (LB-FT) 40 20 20 0 0 0 1000 2000 3000 4000 5000 Engine Speed (RPM)

STP Corrected Power STP Corrected Torque

Figure 3.2.6. STP Corrected Total Average Power and Torque Curves.

The published values for maximum power and torque have been verified. From the curves the maximum power at each speed is now known. The steady-state testing will be conducted on different engine speed intervals for specific powers up to the maximum power for that speed developed in this chapter. The next chapter discusses the steady-state testing and calculates the engine efficiency for each one of the test points.

24 CHAPTER IV

STEADY-STATE ENGINE EFFICIENCY

From the peak power and torque curves a testing strategy was developed for the steady- state efficiency testing. The engine was held at a constant speed and appropriate load to obtain the desired test point. At that point the fuel flow rate was then measured. Using the fuel flow rate the theoretical power was calculated and compared to the measured dynamometer power. From this the engine efficiency at each point was calculated. The uncertainty of the calculations was explored. Finally, the efficiencies were plotted with both the maximum power and torque curves.

4.1 Test Setup

The speed range that the steady-efficiency points were developed over was 1500 to

4000 RPM. This allowed for the engine to be in a region where the turbocharger would be active and is safely below the maximum engine speed. The power ranges for each of the speeds was developed using the peak power curves previously generated. Data was taken in 10 HP intervals for speeds in steps of 250 RPM. This allows for 36 different data points. The number of data points for each speed depended on the maximum engine power as well as the minimum power produced by the engine. At higher engine speeds the minimum engine power increased.

Table 4.1.1 shows the testing speeds.

25 Table 4.1.1. Engine Test Speeds and Power. Speed (RPM) Power (HP) Speed (RPM) Power (HP) Speed (RPM) Power (HP) 1500 10 2250 10 3000 29 1500 20 2250 20 3000 40 1500 30 2250 30 3000 50 1500 MAX 2250 40 3000 60 2250 50 1750 10 2250 60 3500 50 1750 20 3500 60 1750 30 2500 16 3500 70 1750 40 2500 20 3500 80 2500 30 2000 10 2500 40 4000 66 2000 20 2500 50 4000 77 2000 30 2500 60 2000 40 2500 MAX 2000 50

The throttle and load were manipulated to achieve the desired engine speed and power for

the test. The engine is drive by wire so an input voltage could be sent to the engine to keep it at a specific engine speed and power with coordination of the dynamometer load. The dynamometer load was controlled using a dial on the console that adjusted the servomotor controlling the amount of water passing through the brake. The dynamometer was set to a standard test for the steady- state data acquisition. The power and speed were then measured for 72 seconds based on the memory capacity of the dynamometer taking one sample per second. The fuel consumption was measured using diagnostic software from Ross Technologies called VAG-COM. VAG-COM recorded the measured fuel consumption from the engine electronic control unit (ECU). The VAG-

COM sampling rate varies from 1.3 to 4.0 samples per second [15]. The fuel consumption for each test point remained relatively constant (Figs. 4.1.1). Data was taken at points when the engine coolant system reached a stable temperature. Since the tests were run for 72 seconds it was difficult to ensure that all engine test points occurred at the same engine temperature. For certain 26 loads the engine coolant system could not maintain a temperature and data had to be taken as the engine increased in temperature. These loads were at high engine speeds and high power. To try to ensure that the engine temperature was not increasing between test runs the engine was brought back to idle speed. The tests were also run in a random order for each speed for example at 2000 RPM the order of testing was 10 HP, 40 HP, 30 HP, 20 HP, and 50 HP. Several test points were repeated. The throttle input and dynamometer load were also recorded are retested for certain points to make sure that the data would remain consistent from test to test (Table 4.1.2).

The dynamometer load is a percentage of the available load the dynamometer can put on the engine.

5

4

3

2 Fuel Consumption(gal/hr)

1

0 0 500 1000 1500 2000 2500 Engine Speed (RPM)

Figure 4.1.1. Fuel Consumption for Engine Power of 10 HP at 1500 RPM and 50 HP at 2000 RPM.

27 Table 4.1.2. Testing Loads and Throttle Percentages for 1500 RPM. Speed Power Load Throttle RPM HP % % 1500 10 60.0 20.5 1500 20 84.5 42.0 1500 30 89.5 49.5 1500 MAX 92.0 100.0

4.2 Steady-State Efficiency Calculations

The overall efficiency of a diesel engine can be represented as a combined result of a

series of individual efficiencies shown in Equation 4.2.1, where ηe is the overall efficiency, ηth is

the theoretical efficiency, η g is the efficiency of the high-pressure work process, and η m is the mechanical efficiency [1].

ηe = ηth ⋅η g ⋅ηm (4.2.1)

The theoretical efficiency is based on the ideal diesel cycle. It represents the theoretical work in

relation to the energy content of the fuel. The assumptions made for this ideal process are no flow

related losses, ideal gas quality, constant specific heat, and infinite speed of heat propagation and

dissipation. The efficiency of high pressure work process relates the theoretical process and the

high pressure work process. It takes into account the heat and flow related losses of the real gas

exchange phase. Finally, the mechanical efficiency includes the mechanical losses due to friction.

The measured power incorporates the losses due to friction as well as the efficiency of the high

pressure work process since it is being measured not a theoretical calculation.

28 The Carnot efficiency is another calculation of efficiency of a power cycle for a system

operating reversibly through two thermal reservoirs at temperatures TH and TC [3]. The equation

for Carnot efficiency is:

TC η max = 1− (4.2.2) TH

If the operating temperatures of the engine were known then the Carnot efficiency could represent

the best efficiency of that operating point.

From the ideal Diesel cycle as discussed previously the thermal efficiency can be found

as:

k 1  rc −1  η = 1− k −1   (4.2.3) r k()rc −1 

where r is the compression ratio, rc is the cutoff ratio, and k is the specific heat ratio and is

constant in this case [3]. The cutoff ratio can be found if the temperature ratio or volume ratio is

known for the constant pressure heat addition between point 2 and 3 (Figure 1.4) in the Diesel

cycle. Using approximate temperature numbers for points 2 and 3 of 482 °F, the minimum

temperature for diesel fuel to combust, and 752 °F, the cutoff ratio can be found as 1.6. Using a k of 1.4 for air and the compression ratio of 17.5 the thermal efficiency is 64.8%. This is an example of the thermal efficiency for the ideal Diesel cycle.

The efficiency of the Carnot cycle is always higher than the efficiency of the Ideal Diesel

Cycle. This can be seen in the temperature-entropy diagram for the two cycles. Both cycles are internally reversible but the difference lies in the heat transfer from the reservoir to the fuel.

29

Figure 4.2.1. Temperature-Entropy Diagram for Carnot and Ideal Diesel Cycles.

Using the measured torque, speed and the fuel consumption over the 72-second test the engine efficiency can be calculated. First an average of the speed, torque and fuel consumption was taken. Then the theoretical power was calculated from the known higher hearting value of the fuel and the measured fuel consumption. The heating value of a fuel is a number that is equal to the magnitude of the enthalpy of combustion. There are two standard heating values a higher heating value (HHV) and a lower heating value (LHV). The HHV is obtained when all the water formed by combustion is a liquid. The LHV is obtained when all the water formed by combustion is a vapor. The HHV is greater than the LHV by the amount of energy required to vaporize the liquid that is formed [3]. By using the HHV the highest amount of theoretical power or amount of heat can be calculated. Using the LHV would give a lower amount of theoretical power and the resulting engine efficiency would be higher. Standard No. 2 diesel was used for all tests. For standard

No. 2 diesel the higher heating value is 18521 Btu/lb [10]. From this data the efficiency at each speed and power point was calculated (Eq. 4.2.4).

30 Measured Power Tω Efficiency = = (4.2.4) Fuel Consumption × HHV m& f HHV

where T is the engine torque, ω is the engine speed, m& f is the fuel flow rate, and HHV is the higher heating value of No. 2 diesel fuel. The calculated steady-state efficiencies can be found in

Table 4.2.3. The greatest efficiencies occur in the power range at 1750 RPM, with the highest calculated efficiency of 39.6% at a maximum power for 1750 RPM. Efficiencies greater than 34% were plotted on both power and torque versus engine speed maps as points to illustrate the distribution of engine efficiency (Figs. 4.2.1, 4.2.2). The efficiencies were plotted as points on the graphs because there was not enough data to accurately depict the efficiencies as contours. From this data operating points could be determined based on power requirements for various operating conditions. In a conventional vehicle application the road load could be determined based on factors such as desired speed, tire rolling resistance, vehicle weight distribution, and aerodynamics. From that power requirement information Table 4.2.3 could be used to determine the best engine operating speed to obtain the highest efficiency for that power.

31 120.00

100.00

80.00 Peak Power 35% 36% 60.00 37% 38%

Engine Power (HP) 40.00 39%

20.00

0.00 0 500 1000 1500 2000 2500 3000 3500 4000 4500 Engine Speed (RPM)

Figure 4.2.1. Steady-State Efficiency Points plotted for Engine Power.

200.00

180.00 160.00

140.00 Peak Torque 120.00 35% 36% 100.00 37% 80.00 38%

Engine Torque (LB-FT) 60.00 39%

40.00 20.00

0.00 0 500 1000 1500 2000 2500 3000 3500 4000 4500 Engine Speed (RPM)

Figure 4.2.2. Steady-State Efficiency Points plotted for Engine Torque.

32 Figure 4.2.3 shows the optimal efficiency curve from the test data. This shows the highest

efficiency point for each of the tested speeds. If the engine was controlled in a way to ensure the speed and load followed this curve then high fuel efficiency could be realized.

100 90 80 70 60 50 40

Engine Power (HP) 30 20 10 0 0 500 1000 1500 2000 2500 3000 3500 4000 4500 5000 Engine Speed (RPM)

Peak Power Best Efficiency Poly. (Best Efficiency)

Figure 4.2.3. Best Efficiency Curve.

The steady-state efficiencies were plotted versus engine speed for different engine powers

(Figures 4.2.4-9). These figures show how the engine efficiency varies relative to engine speed for

a set engine power. This information could be used to determine an operating point based on

desired engine power.

33 45.0% 40.0% 35.0% 30.0% 25.0% 20.0% 15.0%

Engine Efficiency (%) 10.0% 5.0% 0.0% 1766.3 1538.1 2013.9 2276.1 Engine Speed (RPM)

Figure 4.2.4. Engine Efficiency for 10 HP at Various Engine Speeds.

45.0% 40.0% 35.0% 30.0% 25.0% 20.0% 15.0%

Engine Efficiency (%) 10.0% 5.0% 0.0% 1466.3 1741.7 2022.1 2243.9 2524.4 Engine Speed (RPM)

Figure 4.2.5. Engine Efficiency for 20 HP at Various Engine Speeds.

34 38.5% 38.0% 37.5% 37.0% 36.5% 36.0% 35.5%

Engine Efficiency (%) 35.0% 34.5% 34.0% 1492.5 1483.1 1740.5 1986.0 2274.2 2488.9 Engine Speed (RPM)

Figure 4.2.6. Engine Efficiency for 30 HP at Various Engine Speeds.

40.0% 39.0% 38.0% 37.0% 36.0% 35.0% 34.0%

Engine Efficiency (%) 33.0% 32.0% 31.0% 2024.7 1751.0 2270.8 2544.6 3056.9 Engine Speed (RPM)

Figure 4.2.7. Engine Efficiency for 40 HP at Various Engine Speeds.

35 37.5%

37.0%

36.5%

36.0%

35.5% Engine Efficiency (%) 35.0%

34.5% 2244.9 2489.6 3039.2 3512.1 Engine Speed (RPM)

Figure 4.2.8. Engine Efficiency for 50 HP at Various Engine Speeds.

38.5% 38.0% 37.5% 37.0% 36.5% 36.0% 35.5% 35.0% Engine Efficiency (%) Efficiency Engine 34.5% 34.0% 33.5% 2223.0 2475.7 2982.7 3488.4 Engine Speed (RPM)

Figure 4.2.9. Engine Efficiency for 60 HP at Various Engine Speeds.

Repeated test runs were compared to see if the data was repeatable. Tables 4.2.1 and

4.2.2 show the calculated efficiencies for test run at 1500 RPM and 10 HP as well as 2000 RPM

36 and 20 HP respectively. For 1500 RPM the calculated efficiencies were the same even though the

average engine test speeds were slightly different. For 2000 RPM, the calculated efficiencies only

differed by 0.5%.

Table 4.2.1. Steady-State Efficiency Repeatability for 10 HP at 1500 RPM. Average Average Average Theoretical Test Fuel Flow Actual Actual Actual Heat Input Efficiency Number Rate Speed Torque Power (Power)

# RPM FT-LBS HP gal/hr HP % 1 1538.1 35.1 10.3 0.53 26.6 38.6% 2 1538.3 35.1 10.3 0.53 26.6 38.6%

Table 4.2.2. Steady-State Efficiency Repeatability for 20 HP at 2000 RPM. Average Average Average Theoretical Test Fuel Flow Actual Actual Actual Heat Input Efficiency Number Rate Speed Torque Power (Power)

# RPM FT-LBS HP gal/hr HP % 1 2022.1 54.2 20.9 1.19 59.9 34.8% 2 2015.9 54.8 21.0 1.22 61.2 34.3%

Given a range of engine power and speed requirements for a vehicles operation

Figure 4.2.1 can be used to determine the appropriate operating points using this engine. For

optimal fuel economy in general vehicle applications of this engine the speed and power should

follow the curve seen in Figure 4.2.3.

For a specific application such as series mode for the Akron series-parallel architecture it

would be preferable to run the engine at 1750 RPM. In series mode the power output is limited to the continuous power of the generator. The generator has a peak power of 40 HP and a

continuous power of 28 HP. At 1750 RPM the engine can produce powers up to 40 HP, which

satisfies the maximum power output of the generator. The most efficient point also occurs at the

peak power for 1750 RPM. For higher power requirements, 40 to 60 HP, operating in the range of

37 2250 RPM produces high efficiencies in the range of 37 to 38%. This could help achieve maximum fuel economy at highway speeds of around 75 MPH.

38

Table 4.2.3. Steady-State Efficiencies. Average Average Average Theoretical Fuel Flow Actual Actual Actual Heat Input Efficiency Uncertainty Rate Speed Torque Power (Power) RPM FT-LBS HP gal/hr HP % % 1538.1 35.1 10.3 0.5 26.6 38.6% 5.0% 1466.3 70.3 19.6 1.2 58.6 33.5% 2.3% 1492.5 100.0 28.4 1.6 79.9 35.6% 1.7% 1483.1 114.9 32.4 1.8 90.5 35.8% 1.5%

1766.3 29.9 10.1 0.6 29.3 34.3% 4.6% 1741.7 61.4 20.4 1.1 53.3 38.2% 2.5% 1740.5 88.3 29.3 1.5 77.2 37.9% 1.7% 1751.0 117.0 39.0 2.0 98.5 39.6% 1.4%

2013.9 26.7 10.2 0.7 34.6 29.6% 3.9% 2022.1 54.2 20.9 1.2 59.9 34.8% 2.2% 1986.0 77.2 29.2 1.6 78.5 37.2% 1.7% 2024.7 100.5 38.7 2.1 106.5 36.4% 1.3%

2276.1 27.2 11.8 1.0 50.6 23.3% 2.7% 2243.9 46.0 19.7 1.2 61.2 32.1% 2.2% 2274.2 72.3 31.3 1.6 82.5 37.9% 1.6% 2270.8 93.8 40.6 2.2 109.2 37.2% 1.2% 2244.9 115.7 49.5 2.6 133.1 37.1% 1.0% 2223.0 138.9 58.8 3.1 154.4 38.1% 0.9%

2478.8 33.5 15.8 1.0 47.9 33.0% 2.8% 2524.4 41.2 19.8 1.4 69.2 28.6% 1.9% 2488.9 63.6 30.1 1.6 82.5 36.5% 1.6% 2544.6 85.2 41.3 2.2 109.2 37.8% 1.2% 2489.6 103.2 48.9 2.6 133.1 36.7% 1.0% 2475.7 124.0 58.5 3.1 154.4 37.9% 0.9%

3056.9 70.6 41.1 2.4 119.8 34.3% 1.1% 3039.2 89.7 51.9 2.9 143.8 36.1% 0.9% 2982.7 103.8 58.9 3.3 167.7 35.1% 0.8%

3512.1 73.5 49.2 2.7 138.5 35.5% 1.0% 3488.4 88.7 58.9 3.3 165.1 35.7% 0.8% 3456.3 106.1 69.8 3.8 191.7 36.4% 0.7% 3481.3 118.6 78.6 4.4 223.7 35.1% 0.6%

3976.6 86.7 65.6 4.0 199.7 32.9% 0.7% 3971.1 102.2 77.3 4.7 237.0 32.6% 0.6%

39 4.3 Efficiency Uncertainty

The uncertainty for the measured data must be considered. Error in the data could come

from either noise in the recorded data or from the least significant bit of the measured data.

Figure 4.3.1 shows an example of the recorded data for steady-state testing for 10 HP at

1500 RPM.

15

10

5 Engine Power (HP)

0 1525 1530 1535 1540 1545 1550 1555 Engine Speed (RPM)

Figure 4.3.1. Recorded Engine Power for 1500 RPM.

From the figure it can be seen that the data is relatively constant at 10.3 so the largest error is going to be contributed by the least significant bit of the measured data.

The uncertainty in the efficiencies is calculated using Equation 4.3.1, and is developed using Equations 3.2.1, and 3.2.2 for uncertainty as well as Equation 4.2.4 for efficiency.

40 2 2 2 dη  dT   dω   dm& f  = + +   (4.3.1)       η  T   ω   m& f 

The uncertainty of the efficiency calculations is based on the resolution of the fuel flow rate and

power measurements. The resolution of the fuel flow rate measurement is 0.2 l/hr. At low flow

rates the uncertainty can be as high as 5.0%. At high flow rates the uncertainty drops to 0.6%.

Table 4.2.5 shows that as the flow rate increases the uncertainty decreases.

This method will also be used to calculate the uncertainty in the transient efficiencies in a later chapter.

4.4. Engine Operating Point

The engine coolant temperature was measured by a sensor placed where coolant exits the engine block. Ideally data taken from each test should be at the same engine temperature. For certain speed and power requirements the cooling system was unable to cool the engine and the engine went into thermal runaway. This happened most commonly at high speed and power testing. For lower speed and power requirements the cooling system over cooled the engine. Figure 4.4.1 shows the temperature range and includes the thermal runaway as will as the points that were over cooled. Figure 4.4.2 shows the temperature range without the thermal runaway and over cooled engine points. The engine operating temperature from Figure 4.4.2 varies about 25 °F.

41 250

200

150

100 Engine Temperature (F) 50

0 1000 1500 2000 2500 3000 3500 4000 4500 Engine Speed (RPM)

Figure 4.4.1. Engine Operating Temperature.

250

200

150

100 Engine Temperature (F) 50

0 1000 1500 2000 2500 3000 3500 4000 4500 Engine Speed (RPM)

Figure 4.4.2. Engine Operating Temperature without Thermal Runaway and Overcooling.

42 The steady-state efficiencies give data to help develop an optimal operation point, or curve depending on the vehicle application. In most vehicle applications the ability to maintain a steady- state operating condition is difficult. The transients that are introduced must have some implication on the efficiency and this is explored in the next chapter.

43 CHAPTER V

TRANSIENT ENGINE EFFICIENCY

A constant speed is difficult to obtain in most vehicle situations. With the added complexity of a HEV architecture and multiple operating modes the engine will encounter more transients than a standard vehicle application. Two tests were performed to analyze the effect of varying engine speed on engine efficiency. The first test is a partial throttle test, where only a percentage of full throttle was given to the engine and an acceleration test was performed similar to how the peak power and torque curves were generated. The second test varied the throttle and examined the efficiency when the engine was accelerating and decelerating through various engine speeds. The term throttle is being defined from the sense that the input voltage from the accelerator pedal to the engine electronic control unit was varied. The throttle or accelerator pedal voltage was monitored over its range of motion from unmoved, 0% throttle, to fully depressed 100% throttle. This voltage controls the fuel flow rate of the engine. For a diesel engine the air intake is determined by the engine speed while the power output is controlled by the rate of fuel injected. Partial throttle testing means that a percentage of full voltage was sent to the ECU and the corresponding fuel flow rate was determined. Varied throttle testing means that the input voltage, or fuel flow rate, was varied by an input function was sent to the ECU.

44 5.1 Partial Throttle Testing

The first phase of the transient testing implemented an acceleration test at two different

partial throttles. Tests were conducted at 30% and 40% of full throttle. These tests are similar to

the tests that developed the peak torque and power curves. For 30% of full throttle the lower

engine limit was changed to 1500 RPM and the upper limit was set at 2900 RPM. The range of

engine speeds was decreased due to the limits of the capacity of the dynamometer at lower

speeds. For the 40% of full throttle testing the capacity of the dynamometer was changed and the

engine speed ranged from 1200 RPM to 3400 RPM. The engine was then accelerated at a speed

of 100 RPM per second. The ideal power was again calculated using the measured fuel

consumption and HHV of standard No. 2 diesel fuel. The measured engine power was used to

calculate the efficiency (Tables 5.1.1, 5.1.2). No correction factors were used for the engine

power. The targeted engine speed and actual measured engine speed for the acceleration tests

also matched. Figures 5.1.1 and 5.1.2 show the power curve for the test as well as the efficiency

points for each power and speed for 30% and 40% of full throttle respectively.

The partial throttle transient efficiencies were compared for the different tests. Table 5.1.3 shows the efficiency repeatability comparison for 40% throttle testing results. Efficiencies at engine speeds of 1700, 1800, 2000, and 2100 RPM varied slightly while the rest of the efficiencies were very close. The uncertainty was calculated based on the resolution of the power and fuel flow rate

measurements. Tables 5.1.1 and 5.1.2 also show the uncertainty decreases with an increase in

fuel flow rate.

45 Table 5.1.1. Engine Efficiency and Uncertainty at 30% of Full Throttle. Engine Engine Fuel Efficiency Uncertainty Speed Power Consumption RPM HP gal/hr % % 1500 13.9 0.8 32.6% 3.1% 1600 14.1 1.0 29.4% 2.8% 1700 15.0 1.0 31.3% 2.8% 1800 16.3 1.0 34.0% 2.8% 1900 16.9 1.0 33.4% 2.6% 2000 17.7 1.1 31.7% 2.4% 2100 19.9 1.2 34.0% 2.3% 2200 19.4 1.2 31.7% 2.2% 2300 19.0 1.2 31.0% 2.2% 2400 19.3 1.3 29.0% 2.0% 2500 18.8 1.3 28.2% 2.0% 2600 18.8 1.4 27.2% 1.9% 2700 18.2 1.4 25.3% 1.9% 2800 18.8 1.5 25.2% 1.8% 2900 16.6 1.5 22.3% 1.8%

46 Table 5.1.2. Engine Efficiency and Uncertainty at 40% of Full Throttle. Engine Engine Fuel Efficiency Uncertainty Speed Power Consumption RPM HP gal/hr % % 1200 16.9 0.9 37.3% 3.0% 1300 16.9 1.0 35.3% 2.8% 1400 16.1 1.1 30.2% 2.5% 1500 16.2 1.1 30.4% 2.5% 1600 17.2 1.1 30.8% 2.4% 1700 18.6 1.1 33.3% 2.4% 1800 21.5 1.3 33.6% 2.1% 1900 25.3 1.4 36.5% 1.9% 2000 26.3 1.4 36.6% 1.9% 2100 30.2 1.5 39.1% 1.7% 2200 31.9 1.6 38.6% 1.6% 2300 32.7 1.7 37.2% 1.5% 2400 33.5 1.8 37.0% 1.5% 2500 35.5 2.0 36.0% 1.4% 2600 36.6 2.0 37.2% 1.4% 2700 35.8 2.0 35.4% 1.3% 2800 34.4 2.1 33.1% 1.3% 2900 32.6 2.1 31.4% 1.3% 3000 31.6 2.1 30.4% 1.3% 3100 32.5 2.0 32.1% 1.3% 3200 32.2 2.0 31.8% 1.3% 3300 32.4 2.0 32.9% 1.4% 3400 32.0 1.9 33.4% 1.4%

47 25 40%

35% 20 30%

25% 15

20% Power (HP) 10 Efficiency (%) 15%

10% 5 5%

0 0% 1000 1500 2000 2500 3000 3500 Engine Speed (RPM)

Power Efficiency

Figure 5.1.1. Engine Power and Efficiency at 30% of Full Throttle.

40 45%

35 40%

35% 30 30% 25 25% 20 20% Power (HP) 15 (%)Efficiency 15% 10 10%

5 5%

0 0% 0 500 1000 1500 2000 2500 3000 3500 4000

Engine Speed (RPM)

Power Efficiency

Figure 5.1.2. Engine Power and Efficiency at 40% of Full Throttle. 48 Table 5.1.3. 40% Partial Throttle Efficiency Repeatability.

Engine Test 1 Test 2 Speed Efficiency Efficiency 1400 30.2% 32.1% 1500 30.4% 30.2% 1600 30.8% 31.7% 1700 33.3% 27.2% 1800 33.6% 30.2% 1900 36.5% 35.6% 2000 36.6% 34.6% 2100 39.1% 36.6% 2200 38.6% 38.9% 2300 37.2% 37.4% 2400 37.0% 37.6% 2500 36.0% 36.3% 2600 37.2% 37.9% 2700 35.4% 35.6% 2800 33.1% 33.6% 2900 31.4% 31.0% 3000 30.4% 30.6%

For the 30% of full throttle testing the efficiency ranged from 22.3% at 2900 RPM to 34.0% at 1800 and 2300 RPM. The maximum uncertainty for these efficiencies was 3.1% and occurred at the lowest test speed. For the 40% of full throttle testing the efficiency ranged from 30.2% at

1400 RPM to 39.1% at 2100 RPM. The maximum uncertainty for these efficiencies was 3.0% and also occurred at the lowest test speed. The 40% of full throttle efficiencies seem to follow a curve similar to the power curve for the test. For both partial throttle tests the highest efficiencies occurred in the engine speed range with the maximum torque values.

49 5.2 Varying Throttle Testing

The final transient testing varied the throttle using a sine wave input centered on various

throttle voltages and engine speeds. A total of eleven tests were conducted and data was

recorded for engine speeds of 1500, 1750, 2000 and 2250 RPM. The engine speeds were

narrowed after doing the steady-state testing. In all the tests the engine was either accelerating or decelerating through the engine speed. Figures 5.2.1 and 5.2.2 are two examples of the throttle input to the engine. This testing was also used to help determine the engine response for modeling and control during series operation for another study.

2.000

1.800

1.600

1.400

1.200

1.000

0.800 Throttle Voltage (volts) Voltage Throttle 0.600

0.400

0.200

0.000 0.00 1.00 2.00 3.00 4.00 5.00 6.00 Time (seconds)

Measured Throttle Voltage Input Throttle Voltage

Figure 5.2.1. Throttle Voltage Input for Transient Testing at 1500 RPM.

50 3.500

3.000

2.500

2.000

1.500 Throttle Voltage (volts)

1.000

0.500

0.000 0.00 2.00 4.00 6.00 8.00 10.00 12.00 Time (seconds)

Measured Throttle Voltage Input Throttle Votage

Figure 5.2.2. Throttle Voltage Input for Transient Testing at 1750 RPM.

The engine electronic control unit (ECU) measures voltages in set values so the difference in the two curves on the figures is from the rounding done by the ECU. There was also a difference in the input and measured data due to when the input function was given to the ECU and when the recording of the data started. This appeared as a phase shift in the data. The engine dynamometer did not have the capabilities to record the torque for these tests so all of the data was obtained from the ECU through the use of diagnostic software developed by Ross-

Technologies. The ECU measured torque had to be scaled to the dynamometer torque because

51 the ECU torque is a calculated torque based on fuel consumption and is more ideal. Information

about how the ECU specifically calculates the engine torque is not available. In the diagnostic

software engine torque is grouped with engine speed and fuel flow rate. From this either a look up

table from previously known torque values or a calculation could be used to determine the engine

torque. The steady state data was used to develop the relationship between the dynamometer measured torque and the ECU estimated torque for each speed. A linear trendline was used to develop this relationship since most of the graphs appeared to be linear. The equation for each speed was then used with the recorded data to calculate the power for each speed and fuel consumption rate. Figure 5.2.3 shows the torque relationship used for 2000 RPM.

140.0

120.0

100.0

80.0

60.0 y = 1.0463x + 11.352 ECU Torque (LB-FT) Torque ECU R2 = 0.9979 40.0

20.0

0.0 0.0 20.0 40.0 60.0 80.0 100.0 120.0 Measured Torque (LB-FT)

Figure 5.2.3. Torque Relationship for 2000 RPM.

52 The efficiency was calculated for each speed and fuel consumption rate (Equation 4.3.1).

The below figures show efficiency points plotted versus the engine power for each of the engine speeds. Error bars are placed on each data point. The uncertainty was again based on the resolution of the measured fuel flow rate. The efficiencies for 1500 RPM ranged from 30.3% to

38.5%. The efficiencies for 1750 RPM ranged lower from 15.5% to 29.9%. For 2000 RPM the efficiencies ranged from 17.8% to 40.4%. The efficiency point of 40.4% occurred at a higher speed and load condition than what was tested in the steady-state data. Finally, the efficiency range for

2250 RPM was 25.6% to 35.7%. For 1500 RPM the efficiencies seem relatively constant with increasing power. For 1750, 2000, and 2250 RPM the efficiencies increase with increasing power.

53 45.0%

40.0%

35.0%

30.0%

25.0%

20.0% Efficiency (%) Efficiency

15.0%

10.0%

5.0%

0.0% 0.0 5.0 10.0 15.0 20.0 25.0 Power (HP)

Figure 5.2.4. Transient Efficiencies for 1500 RPM.

54 35.0%

30.0%

25.0%

20.0%

15.0% Efficiency (%) Efficiency

10.0%

5.0%

0.0% 0.0 5.0 10.0 15.0 20.0 25.0 30.0 35.0 40.0 Power (HP)

Figure 5.2.5. Transient Efficiencies for 1750 RPM.

55 45.0%

40.0%

35.0%

30.0%

25.0%

20.0% Efficiency (%) Efficiency

15.0%

10.0%

5.0%

0.0% 0.0 5.0 10.0 15.0 20.0 25.0 30.0 35.0 40.0 45.0 50.0 Power (HP)

Figure 5.2.6. Transient Efficiencies for 2000 RPM.

56 40.0%

35.0%

30.0%

25.0%

20.0% Efficiency (%) Efficiency 15.0%

10.0%

5.0%

0.0% 0.0 5.0 10.0 15.0 20.0 25.0 30.0 35.0 40.0 Power (HP)

Figure 5.2.7. Transient Efficiencies for 2250 RPM.

Transient efficiencies for both partial and varied throttles have been calculated. From the partial throttle testing it appears that during acceleration the efficiencies are highest in the peak torque region of engine speed. For the varied throttle testing around 1500 RPM the efficiencies seem to remain relatively constant. This could be because the turbocharger is not very active in this speed range. For the other speeds tested in the varied throttle testing, 1750, 2000, and

2250 RPM, the efficiencies increase as the engine power increases. Now that both steady-state

57 and transient efficiencies have been determined from several different test procedures they can now be compared. The next chapter discusses differences and similarities between the steady- state results and the transient results.

58 CHAPTER VI

STEADY-STATE AND TRANSIENT ENGINE EFFICENCY COMPARISON

Now that an analysis for both steady-state and transient efficiency has been completed a comparison and conclusions must be developed. First the comparison between the steady-state and partial throttle testing is examined. The appropriate efficiency points are compared in relation to engine speed and power. Then the comparison between the steady-state and varied throttle testing is examined. For the varied throttle comparison multiple test points are compared to steady-state data. Finally the effects of varied throttle testing on vehicle fuel efficiency are explored.

6.1 Steady-State and Partial Throttle Comparisons

The efficiencies taken during the partial throttle testing represent an instantaneous velocity at that point as it is accelerating at 100-RPM per second. Since the engine is accelerating the efficiencies are expected to be lower than the steady-state values. The steady-state efficiency and partial throttle transient efficiencies were plotted for 1500 RPM, 2000 RPM and 2500 RPM.

59 39.0%

38.0% 10.3 HP 37.0%

36.0%

35.0% Efficiency (%) Efficiency 34.0% 19.6 HP 33.0% 13.9 HP 32.0% 1460 1470 1480 1490 1500 1510 1520 1530 1540 1550 Engine Speed (RPM)

Throttle 30% Steady-State

Figure 6.1.1. Steady-State and 30% Partial Throttle Comparison for 1500 RPM.

36.0% 20.9 HP 35.0%

34.0%

33.0%

32.0% 17.7 HP Efficiency (%) Efficiency 31.0%

30.0% 10.2 HP 29.0% 1995 2000 2005 2010 2015 2020 2025 Engine Speed (RPM)

Throttle 30% Steady-State

Figure 6.1.2. Steady-State and 30% Partial Throttle Comparison for 2000 RPM.

60 34.0% 15.8 HP 33.0%

32.0%

31.0%

Efficiency (%) Efficiency 30.0% 19.8 HP 29.0% 18.8 HP 28.0% 2470 2480 2490 2500 2510 2520 2530 Engine Speed (RPM)

Throttle 30% Steady-State

Figure 6.1.3. Steady-State and 30% Partial Throttle Comparison for 2500 RPM.

45.0% 40.0% 19.6 HP 10.3 HP 35.0% 30.0% 16.2 HP 25.0% 20.0%

Efficiency (%) Efficiency 15.0% 10.0% 5.0% 0.0% 1460 1480 1500 1520 1540 1560 Engine Speed (RPM)

Throttle 40% Steady-State

Figure 6.1.4. Steady-State and 40% Partial Throttle Comparison for 1500 RPM.

61 37.5% 29.2 HP 37.0% 26.3 HP 36.5%

36.0%

Efficiency (%) Efficiency 35.5%

35.0% 20.9 HP 34.5% 1980 1990 2000 2010 2020 2030 Engine Speed (RPM)

Throttle 40% Steady-State

Figure 6.1.5. Steady-State and 40% Partial Throttle Comparison for 2000 RPM.

38.0% 41.3 HP 37.8% 37.6% 37.4% 37.2% 37.0% 36.8%

Efficiency (%) Efficiency 36.6% 36.4% 30.1 HP 36.2% 35.5 HP 36.0% 35.8% 2480 2490 2500 2510 2520 2530 2540 2550 Engine Speed (RPM)

Throttle 40% Steady-State

Figure 6.1.6. Steady-State and 40% Partial Throttle Comparison for 2500 RPM.

62 In Figure 6.1.1 the transient efficiency was calculated at 13.9 HP while the steady-state values shown are for powers of 10.3 HP and 19.6 HP. The transient efficiency of this point lies within the

boundaries of the steady-state data. Figure 6.1.2 shows that the steady-state data was taken at

slightly higher speeds than the transient data. The measured power at this point is 17.7 HP and

the efficiency closely matches that of the steady-state efficiency for 20.9 HP. This is also the case

in Figure 6.1.3 where the transient efficiency is for 18.8 HP and it is close to the steady-state

efficiency for 19.8 HP. Similar trends can be seen in Figures 6.1.4, 6.1.4 and 6.1.6. The fuel

consumption for each one of these points is also within the range of the steady-state data. Since

the 40% throttle testing has a higher fuel consumption rate than the 30% testing there can not be a

comparison between the two. Possibly by changing the rate of engine acceleration some common

points between the two partial throttle tests could be found.

6.2 Steady-State and Varied Throttle Comparisons

The varied throttle efficiency for engine speeds of 1500, 1750, 2000, and 2250 RPM was

compared to the appropriate steady-state data. The throttle was varied at a rate of 0.5 Hz. The

efficiencies were then compared based on engine power at that point. Figures 6.2.1, 6.2.2, 6.2.3,

and 6.2.4 show the steady-state efficiency curve and the transient efficiency curve for each of the

speeds.

63 39.0%

38.0%

37.0%

36.0%

35.0% Efficiency (%) Efficiency 34.0%

33.0%

32.0% 0.0 5.0 10.0 15.0 20.0 25.0 30.0 35.0 Engine Power (HP)

Steady-State Transient

Figure 6.2.1. Steady-State and Varied Throttle Efficiency Comparison for 1500 RPM.

45.0% 40.0% 35.0% 30.0% 25.0% 20.0%

Efficiency (%) Efficiency 15.0% 10.0% 5.0% 0.0% 0.0 5.0 10.0 15.0 20.0 25.0 30.0 35.0 40.0 45.0 Engine Power (HP)

Steady-State Transient

Figure 6.2.2. Steady-State and Varied Throttle Efficiency Comparison for 1750 RPM.

64 40.0% 35.0% 30.0% 25.0% 20.0% 15.0% Efficiency (%) Efficiency 10.0% 5.0% 0.0% 0.0 5.0 10.0 15.0 20.0 25.0 30.0 35.0 40.0 45.0 Engine Power (HP)

Steady-State Transient

Figure 6.2.3. Steady-State and Varied Throttle Efficiency Comparison for 2000 RPM.

40.0% 35.0% 30.0% 25.0% 20.0% 15.0% Efficiency (%) Efficiency 10.0% 5.0% 0.0% 0.0 10.0 20.0 30.0 40.0 50.0 60.0 70.0 Engine Power (HP)

Steady-State Transient

Figure 6.2.4. Steady-State and Varied Throttle Efficiency Comparison for 2250 RPM.

65 For each engine speed the transient efficiency curve followed the steady-state curve but

just at a lower efficiency. For 1500 RPM the transient efficiency points for 16.6 and 22.4 HP

intersect the steady-state efficiency line. The transient efficiency points for 16.1, 21.2, and 21.6 HP are much lower than the steady-state data. Consideration was taken to determine if there was an

effect from the engine accelerating or decelerating through the point and no relationship could be

developed.

For the above tests the rotational inertia of the engine is assumed to be mush greater than

the rotational inertia of the dynamometer. The rotational inertia of the engine and dynamometer

could affect the results of the transient tests. The polar moment of inertia of the engine and

dynamometer break should be compared. Qualitatively the polar moment of inertia of the engine

would include the crankshaft, turbocharger, pistons, connecting rods, overhead camshaft, and

flywheel. The combination of all of these items would be greater that the internal components of

the dynamometer. If this was not the case then there would have also been errors in the

acceleration testing for peak power and torque. In addition, for actual application of the data a

transmission will be added to the system, which would introduce additional rotational inertia.

6.3 Steady-State and Transient Fuel Economy Comparison

For another point of comparison vehicle fuel economy can be examined. Vehicle fuel

economy can be related to engine efficiency through a power balance as seen in Equation 6.2.1.

FRL ⋅ v Q& f ⋅ HHV ⋅η E = (6.2.1) η DT

66 In this equation Q& f is the volumetric fuel flow rate, HHV is the higher heating value of the fuel,

η E is the engine efficiency, FRL is the road load force, v is the vehicle speed, and η DT is the drivetrain efficiency. By rearranging terms miles per gallon or fuel economy can be calculated as:

v η ⋅η ⋅ HHV = E DT (6.2.2) Q& f FRL

Vehicle speed has units of distance over time, while volumetric fuel flow rate has unit of volume

over time. By dividing the two terms the result is distance over volume. By using the right

dimensions for the variables in Equation 6.2.1 miles per gallon can be determined based on a

steady speed. From Equation 6.2.2 it can be seen that fuel economy is directly proportional to

engine efficiency. If engine efficiency decreases then fuel economy also decreases.

From this initial transient testing it appears that the transient efficiency follows the same

curve as the steady-state data. The partial throttle testing showed that the change in efficiency

was still within the reasonable range developed by the steady-state data.

67 CHAPTER VII

SUMMARY

The objective of the project was to develop the steady-state engine efficiencies for a 2005

VW 1.9 L diesel engine and consider transient effects on these efficiencies. The work consisted of research of past, present, and future development of diesel engines. The engine researched was selected based on the requirements and research conducted for a national advanced vehicle design competition, Challenge X. Acceleration testing was performed to validate published data for the selected engine. The acceleration testing was also used to develop the engine operation points for steady-state testing. The fuel flow rate measured from steady-state testing was used to develop engine efficiencies. These engine efficiencies were related to vehicle fuel economy for steady speeds. Partial throttle and varied throttle testing was conducted to study transient effects on diesel engine efficiency. The transient efficiencies were compared to the steady-state efficiencies. The varying throttle transient testing showed that the efficiencies do decrease with an accelerating/decelerating engine. Initial transient tests also show that the efficiencies decrease but maintain a similar curve to the steady-state data. The transient effects on efficiency and how they impacted vehicle fuel economy for steady speeds was also investigated. Finally, suggested operating points for the engine implementation in the series-parallel HEV developed by the

University of Akron were made.

68 The technical contributions of this work are listed as follows:

1. The published values for peak engine power and torque for the 2005 1.9L VW engine have

been verified.

2. Steady-state efficiency points for the 2005 1.9L VW engine have been developed.

3. The engine efficiency and vehicle fuel economy for a steady speed have been related.

4. The effects of partial throttle acceleration testing on engine efficiency of the 2005 1.9L VW

engine was explored and compared to steady-state results.

5. The effects of varied throttle testing on engine efficiency of the 2005 1.9L VW engine was

explored and compared to steady-state results.

6. Transient engine efficiency loss and its effect of vehicle fuel economy has been shown.

Future work for this project can consist of the following:

1. Further steady-state testing at different engine power and speeds to develop efficiency

contours.

2. Additional partial throttle testing at lower percentages of throttle.

3. Partial throttle testing through the engine speed range with peak torque values to see if the

efficiency remains high in that speed range.

4. Varied throttle testing at different voltage input frequencies to represent driving response.

5. Exploration into the effects of using alternative fuels such as B20 biodiesel on the engines

performance in both steady-state and transient operation.

6. Calculations and measurements of the engines polar moment of inertia both theoretically

and with the use of a dynamometer.

69 REFERENCES

[1] R. Bosch, “Diesel-Engine Management”, SAE International, Warrendale, PA, 2004.

[2] C. Cowland, “Passenger Vehicle Diesel Engines for the U.S.”, SAE International, Warrendale, PA, 2004.

[3] M. Moran, “Fundamentals of Engineering Thermodynamics”, John Wiley & Sons, INC., New York, 2000.

[4] F. An, “Evaluating Commercial and Prototype HEVs”, Argonne National Laboratory, 2001.

[5] R. Rowe, “Design and Optimization of the University of Wisconsin’s Parallel Hybrid Electric Sport Utility Vehicle” SAE International, Warrendale, PA, 2002.

[6] R. Krieger, “Diesel Engines: One Option to Power Future Personal Transportation Vehicles”, GM Research and Development Center, SAE International, Warrendale, PA, 1997.

[7] J. Marshaus, “Design and Testing of a Prototype Midsize Parallel Hybrid Electric Sport Utility Vehicle”, SAE International, Warrendale, PA, 2004.

[8] S. Burch, “Trading off HEV Fuel Economy and Emissions through Optimization”, National Renewable Energy Laboratory, 1999.

[9] C. D. Rakopoulos, “The Effect of Various Dynamic, Thermodynamic and Design Parameters on the Performance of a Turbocharged Diesel Engine Operating under Transient Load Conditions”, SAE International, Warrendale, PA, 2004.

[10] D. Holt, “Alternative Diesel Fuels”, SAE International, Warrendale, PA, 2004.

[11] Y. Wang, “Performance and Emissions Characteristics of a Diesel Engine Using Vegetable Oil Blended Fuel”, ASME, 2005.

[12] D. Kawano, “Applications of Biodiesel Fuel to Modern Diesel Engine”, SAE International, Warrendale, PA, 2006.

70 [13] “WinDyn SF-901 Operator’s Manual”, SuperFlow Computerized Engine and Vehicle Test Systems, 1998.

[14] “1998-2004 Volkswagen New Beetle Service Manual”, Bentley Publishers, Cambridge, MA, 2004.

[15] “VAG-COM Diagnostic Software for VW/Audi/SEAT/Skoda User’s Manual Release 512.1”, Ross-Tech LLC, Lansdale, PA, 2006.

71

APPENDICIES

72 APPENDIX A

VOLKSWAGEN ENGINE DATA

Displacement liters 1.9 Output HP at RPM 100/4000 LB-FT at 177/1800 to Torque RPM 2400 Bore Diameter in 3.13 Stroke mm 3.76 Compression ratio 17.5 Ignition sequence 1-3-4-2 Catalytic converter Exhaust gas recirculation Turbocharging Charge air cooling

73 APPENDIX B

EXPERIMENTAL TEST SETUP

Figure B.1. Modified Transmission Bell Housing.

74

Figure B.2. Modified Flywheel.

Figure B.3. Custom Bell Housing Extension.

75

Figure B.4. Modified Innovations Engineering Torque Dampener.

Figure B.5. Innovations Engineering Dynamometer Adapter.

76 APPENDIX C

CONTACT INFORMATION

The following contact information is provided for additional questions. The recorded data is also available through these contacts.

Mr. Charles Van Horn

Email: [email protected]

Dr. Scott Sawyer

Email: [email protected]

Dr. Richard Gross

Email: [email protected]

77