Gaseous and Particulate Matter Emissions of a Supercharged ... · that the current EPA regulations are 99.4 mg*NOx/km, this engine configuration represents a possible replacement

Gaseous and Particulate Matter Emissions of a

Supercharged Spark Ignited Hydrogen Fueled Internal

Combustion Engine

by

Sean Kieran

A thesis submitted in conformity with the requirements

for the degree of Master of Applied Science

Mechanical and Industrial Engineering

University of Toronto

© Copyright by Sean Kieran 2016

ii

Gaseous and Particulate Matter Emissions of a Supercharged

Spark Ignited Hydrogen Fueled Internal Combustion Engine

Sean Kieran

Master of Applied Science

Mechanical and Industrial Engineering

University of Toronto

2016

Abstract

A spark ignited hydrogen fueled engine was operated at three equivalence ratios (0.4, 0.5, and

0.6) with a supercharger. During steady-state road load conditions, the engine produced

exceptionally low unburned hydrocarbon, carbon monoxide, carbon dioxide, and particulate

matter emissions. The oxides of nitrogen (NOx) emissions of the supercharged engine were 31.4,

149.5, and 787.0 mg*NOx/km for the equivalence ratios 0.4, 0.5, and 0.6 respectively. Given

that the current EPA regulations are 99.4 mg*NOx/km, this engine configuration represents a

possible replacement option for gasoline fueled engines without the need for exhaust

aftertreatment. During engine start-up, some of the supercharged tests exhibited particulate

matter emission spikes. These particulate matter spikes do not seem to be related to equivalence

ratio, coolant temperature, testing order, or start-up acceleration. Currently, there is no

explanation why some of the tests produced particulate matter during engine start-up and others

did not.

iii

Acknowledgments

I would like to first thank Dr. Wallace for giving me the opportunity to do this research. He is

truly an expert in this field and I feel very privileged to have learned from him. Dr. Wallace

struck a perfect balance between giving me the freedom to research what interested me and

giving me the guidance to do it. I have learned a lot from Dr. Wallace and I hope to take his

supervisory techniques into my own career.

There was always an atmosphere of comradeship at the Engine Research and

Development Lab (ERDL). Over the course of this degree I was fortunate to have worked with

several gifted people. Alin Pop and I worked to setup this engine; first running it on natural gas.

Working through difficult problems together and facing adversity on the project definitely

brought us closer together. I will forever be in debt to Alin for his hard work both before my time

at the University of Toronto and during.

I am also grateful to have worked with Ivan Gogolev, Khaled Rais, Bryden Smallwood,

Abbas Ali, Kang Pan, Manuel Ramos, and Dan Chown. Much of the work in this thesis would

not have been possible without their advice and input. Another group that I owe a debt of

gratitude is the technical staff in Mechanical and Industrial Engineering. Osmond Sargeant,

Terry Zakk, Tony Ruberto, and the MC78 machine shop staff all contributed invaluable help to

the project.

Finally, I would like to thank my family. Without their love and support, none of this

would have been possible.

iv

Table of Contents

1. Introduction ............................................................................................................................... 1

Origin of Test Engine.......................................................................................................... 3 1.1.

2. Background ............................................................................................................................... 4

Unique Properties of Hydrogen .......................................................................................... 4 2.1.

Hydrogen Fueled Internal Combustion Engines ................................................................. 5 2.2.

2.2.1. Abnormal Combustion ............................................................................................ 5

2.2.2. Design Considerations ............................................................................................ 7

2.2.3. Operating Strategies ................................................................................................ 9

2.2.4. Emissions .............................................................................................................. 11

2.2.5. Exhaust Gas Recirculation .................................................................................... 13

2.2.6. Performance .......................................................................................................... 14

Oil Consumption ............................................................................................................... 16 2.3.

Engine Particulate Emissions ............................................................................................ 19 2.4.

2.4.1. Particulate Emissions from Lubricating Oil ......................................................... 22

Human Health Effects of Particulate Emissions ............................................................... 25 2.5.

3. Experimental Setup ................................................................................................................. 26

Naturally Aspirated Engine Configuration ....................................................................... 26 3.1.

Supercharged Engine Configuration ................................................................................. 28 3.2.

Positive Crankcase Ventilation ......................................................................................... 30 3.3.

3.3.1. Oil Coalescing Filter ............................................................................................. 33

Switching Between Supercharged and Naturally Aspirated Configurations .................... 34 3.4.

Exhaust Emissions Equipment .......................................................................................... 34 3.5.

3.5.1. Isokinetic Probe .................................................................................................... 35

3.5.2. Engine Exhaust Particle Sizer ............................................................................... 38

v

3.5.3. Fourier Transform Infrared Spectroscopy ............................................................ 41

3.5.4. Emissions Bench ................................................................................................... 42

3.5.5. LICOR 840A ......................................................................................................... 43

3.5.6. AFRecorder 2400 .................................................................................................. 44

Data Acquisition ............................................................................................................... 44 3.6.

Throttle Body Controller................................................................................................... 46 3.7.

Dynamometer .................................................................................................................... 46 3.8.

Electronic Control Unit ..................................................................................................... 47 3.9.

4. Methodology and Experimental Procedure ............................................................................ 48

Test Matrix ........................................................................................................................ 48 4.1.

General Test Protocol ....................................................................................................... 49 4.2.

Specific Testing Protocol for Spark Timing Tests ............................................................ 50 4.3.

Specific Testing Protocol for Naturally Aspirated Tests .................................................. 51 4.4.

Specific Testing Protocol for Supercharged Tests ............................................................ 52 4.5.

Filter Sample Weights....................................................................................................... 52 4.6.

5. Results ..................................................................................................................................... 53

Spark Timing Tests ........................................................................................................... 54 5.1.

Steady State Tests ............................................................................................................. 60 5.2.

Acceleration During Start-up ............................................................................................ 71 5.3.

Lubricating Oil Consumption Rate ................................................................................... 78 5.4.

Filter Analysis ................................................................................................................... 81 5.5.

Emissions Equipment........................................................................................................ 84 5.6.

6. Discussion and Conclusion ..................................................................................................... 87

7. Future Work ............................................................................................................................ 92

8. References ............................................................................................................................... 95

vi

9. Appendices ............................................................................................................................ 100

Conversion from SL/min of Hydrogen to g/s ................................................................. 100 9.1.

Equivalence Ratio Calculations ...................................................................................... 100 9.2.

Lubricating Oil Consumption Rate Calculations ............................................................ 102 9.3.

Converting Emissions to a Per km Basis ........................................................................ 104 9.4.

LICOR CO2 Measurement Correction ........................................................................... 104 9.5.

Fuel Contribution to CO and CO2 .................................................................................. 106 9.6.

Available Turbocharger Power ....................................................................................... 108 9.7.

Road Load Power ............................................................................................................ 109 9.8.

Supercharger Power Calculations ................................................................................... 110 9.9.

Throttle Body Arduino Code .......................................................................................... 112 9.10.

Clean Room Procedure ................................................................................................... 115 9.11.

Emissions Operating Procedure ...................................................................................... 116 9.12.

Calibration Procedure for Sensors .................................................................................. 124 9.13.

9.13.1. Mass Air Flow Sensor ......................................................................................... 124

9.13.2. Pressure Sensors.................................................................................................. 131

9.13.3. Temperature Sensors ........................................................................................... 143

Filter Elements ................................................................................................................ 160 9.14.

Minimum Dilution Ratio Calculations............................................................................ 167 9.15.

PM Correction to Raw Exhaust Gas Basis ..................................................................... 172 9.16.

vii

List of Tables

Table 1. Engine Specifications ..................................................................................................... 27

Table 2. EEPS Dilution Settings ................................................................................................... 40

Table 3. Emissions Bench Channels and Ranges ......................................................................... 43

Table 4. Test Matrix ...................................................................................................................... 48

Table 5. Variables for Road Load Power Calculation ................................................................ 109

Table 6. Time between Actuation of Micro Switches and Voltage Output of MAF Sensor from

Test on July 31st, 2014 ................................................................................................................ 127


Test on August 11th

, 2014 ........................................................................................................... 127

Table 8. MAF Sensor Output Voltage and Calculated Mass Air Flow Rate from Tests on July

31st, 2014 and August 11

th, 2014 ................................................................................................ 128

Table 9. Pressure and MAP Sensor #1 Voltage Output from Test on August 12th

, 2014 .......... 134

Table 10. Pressure Converted to kPa and MAP Sensor #1 Voltage Output ............................... 134


, 2014 ........ 135



, 2014 ........ 137



, 2014 ........ 138


Table 17. Fuel Rail Pressure Sensor Voltage Output and Pressure from Test on August 13th

, 2014

..................................................................................................................................................... 140

Table 18. Fuel Rail Pressure Sensor Voltage Output and Pressure Converted to kPa ............... 141

Table 19. Temperature and Resistance of MAP Sensor #1 from Test on August 8th

, 2014 ....... 147


, 2014 ..... 147

Table 21. Temperature and Resistance of MAP sensor #2 from Test on August 8th

, 2014 ........ 148


, 2014 ..... 148


, 2014 ....... 149


, 2014 ..... 149


, 2014 ..... 150

viii

Table 26. Temperature and Resistance of MAF Sensor from Test on August 18th

, 2014 .......... 150


, 2014 ..... 151

Table 28. Resistance of Various Temperature Sensors at Zero Degrees from Test on August 19th

,

2014............................................................................................................................................. 151

Table 29. Temperature and Resistance of Coolant Temperature Sensor from Test on August 15th

,

2014............................................................................................................................................. 152


,

2014............................................................................................................................................. 152

Table 31. Temperature and Resistance of Oil Temperature Sensor from Test on August 15th

,

2014............................................................................................................................................. 153


,

2014............................................................................................................................................. 153

Table 33. Temperature and Resistance of Fuel Rail Temperature Sensor from Test on August

20th

, 2014 .................................................................................................................................... 154


21st, 2014 ..................................................................................................................................... 154

Table 35. Dilution Ratio Calculation for an Equivalence Ratio of 0.6 ....................................... 169

Table 36. Dilution Ratio Calculation for an Equivalence Ratio of 0.4 ....................................... 169

Table 37. Dilution Ratio for an Equivalence Ratio of 0.5 .......................................................... 170

ix

List of Figures

Figure 1. Oil Consumption Mechanisms (Froelund & Yilmaz, 2004) ......................................... 16

Figure 2. Lubricating Oil Consumption Mechanisms from the Piston-Ring-Liner System

(Froelund & Yilmaz, 2004)........................................................................................................... 17

Figure 3. Current Understanding of the Structure of a Complex Engine Exhaust Particle (Matti

Maricq, 2007) ................................................................................................................................ 20

Figure 4. Condition of Diesel Particulate Filter after Regeneration (Givens, et al., 2003) .......... 22

Figure 5. Diagram of Naturally Aspirated Engine Configuration ................................................ 26

Figure 6. Picture of Naturally Aspirated Engine Configuration ................................................... 28

Figure 7. Diagram of Supercharged Engine Configuration .......................................................... 28

Figure 8. Picture of Supercharged Engine Configuration ............................................................. 30

Figure 9. Naturally Aspirated PCV System (G2IC Turbo Guide, 2016) ...................................... 31

Figure 10. PCV System Diagram for Supercharged Configuration (Natkin, et al., 2003) ........... 32

Figure 11. Oil Coalescing Filter (MANN+HUMMEL ProVent, 2016) ....................................... 33

Figure 12. Isokinetic Probe Diagram ............................................................................................ 35

Figure 13. Isokinetic Probe ........................................................................................................... 36

Figure 14. Isokinetic Sampling Flowchart Diagram ..................................................................... 37

Figure 15. Pump (Left) and Diluter (Right) Configuration for EEPS (Matter Engineering, 2014)

....................................................................................................................................................... 39

Figure 16. NOx and Fuel Conversion Efficiency vs. Ignition Timing for φ = 0.4 ....................... 55



Figure 19. Available Turbocharger Power vs. Engine Power for Spark Timing Tests ................ 59

Figure 20. NOx vs. Equivalence Ratio at the Road Load Setting ................................................. 61

Figure 21. NOx Produced per km vs. Equivalence Ratio at the Road Load Setting with Emissions

Regulation Comparisons (United States Environmental Protection Agency, 2014) (Johnson,

2014) (MECA, 2014) .................................................................................................................... 62

Figure 22. Fuel Conversion Efficiency vs. Equivalence Ratio for Supercharged and Naturally

Aspirated Tests at the Road Load Power Setting.......................................................................... 64

Figure 23. Percentage of NO or NO₂ that Contributes to NOx vs. Equivalence Ratio ................ 68

x

Figure 24. Intake Manifold Pressure vs. Engine Power for Supercharged and Naturally Aspirated

Tests with Different Equivalence Ratios ...................................................................................... 70

Figure 25. 1-Minute PM Average Concentration and Engine Speed vs. Time for Supercharged

Spark Timing Test φ = 0.4 February 11, 2016.............................................................................. 72

Figure 26. 1-Minute PM Average Concentration and Engine Speed vs. Time for Supercharged φ

= 0.4 March 10, 2016 .................................................................................................................... 73

Figure 27. Peak 1-Minute PM Average Concentration vs. Peak Engine Acceleration ................ 74

Figure 28. Peak 1-Minute Average PM Average Concentration vs. Coolant Temperature ......... 75

Figure 29. Peak 1-Minute PM Average Concentration vs. Testing Order of that Day................. 76

Figure 30. Peak 1-Minute PM Average Concentration vs. Nominal Equivalence Ratio ............. 77

Figure 31. Lubricating Oil Consumption Rate ............................................................................. 79

Figure 32. Lubricating Oil Consumption Rates of Various Engine Types (Kapetanovic, Wallace,

& Evans, 2009) (Froelund, Menezes, Johnson, & Rein, 2001) .................................................... 80

Figure 33. Mass Collected On Filters ........................................................................................... 81

Figure 34. Clean Filter (Left) and Tested Filter (Right) Naturally Aspirated at an Equivalence

Ratio of 0.4.................................................................................................................................... 82

Figure 35. FTIR NOx vs. Emissions Bench NOx ......................................................................... 84

Figure 36. Water Concentration Measured vs. Theoretical .......................................................... 85

Figure 37. Willan’s Line for Supercharged and Naturally Aspirated Tests ............................... 111

Figure 38. MAF Sensor Calibration Configuration with Bell Prover ......................................... 125

Figure 39. Graph of Mass Air Flow Rate (kg/hr) vs. MAF Sensor Output Voltage (V) ............ 129

Figure 40. MAP Pressure Sensor Calibration Configuration...................................................... 131

Figure 41. Fuel Rail Pressure Sensor Configuration .................................................................. 131

Figure 42. Graph of Pressure vs. MAP Sensor #1 Voltage Output ............................................ 135




Figure 46. Graph of Pressure vs. Fuel Rail Pressure Sensor Voltage Output ............................. 142

Figure 47. Air Temperature Sensor Configuration ..................................................................... 143

Figure 48. Ice Bucket Configuration .......................................................................................... 144

Figure 49. Heated Engine Coolant Bath Configuration .............................................................. 144

xi

Figure 50. Graph of Temperature vs. Resistance of MAP Sensor #1 ......................................... 155




Figure 54. Graph of Temperature vs. Resistance of Fuel Rail Temperature Sensor .................. 157

Figure 55. Graph of Temperature vs. Resistance of MAF Sensor .............................................. 157

Figure 56. Graph of Temperature vs. Resistance of Oil Temperature Sensor ............................ 158

Figure 57. Graph of Temperature vs. Resistance of Coolant Temperature Sensor .................... 158

Figure 58. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio

of 0.4 on February 26th

, 2016 on the Right ................................................................................. 160



, 2016 on the Right ................................................................................. 160



, 2016 on the Right ................................................................................. 161


of 0.5 on March 3rd

, 2016 on the Right ....................................................................................... 161


of 0.6 on March 3rd

, 2016 on the Right ....................................................................................... 162


of 0.4 on March 3rd

, 2016 on the Right ....................................................................................... 162


of 0.5 on March 3rd

, 2016 on the Right ....................................................................................... 163

Figure 65. Unused Filter on the Left and 30 Minute Naturally Aspirated Test at an Equivalence

Ratio of 0.6 on March 5th

, 2016 on the Right ............................................................................. 163



, 2016 on the Right ............................................................................. 164



, 2016 on the Right ............................................................................. 164


of 0.6 on March 10th

, 2016 on the Right ..................................................................................... 165

xii



, 2016 on the Right ..................................................................................... 165

Figure 70. Unused Filter on the Left and 30 Minute Oven Test on the Right ............................ 166

Figure 71. Unused Filter on the Left and 30 Minute Isokinetic Test on the Right ..................... 166

Figure 72. Diagram of Dilution Streams..................................................................................... 167

Figure 73. Psychrometric Chart (Moran, Shapiro, Boettner, & Bailey, 2011) ........................... 171

xiii

Nomenclature

Symbols

φ – fuel/air equivalence ratio

λ – air/fuel equivalence ratio

Abbreviations

°BTDC – degrees before top dead center

C3H8 – propane

CAI – California Analytical Instruments

CI – compression ignition

CO – carbon monoxide

CO₂ – carbon dioxide

CT – EEPS coefficient based on the primary dilution temperature

DI – direct injection

DP – EEPS primary dilution factor

DPF – diesel particulate filter

DR – dilution ratio

DS – EEPS secondary dilution factor

ECM – Engine Control and Monitoring

ECU – electronic control unit

EEPS – engine exhaust particle sizer

xiv

EGR – exhaust gas recirculation

EPA – Environmental Protection Agency

ERDL – Engine Research and Development Lab

FEAD – front end accessory drive

FTIR – Fourier transform infrared spectroscopy

GDI – gasoline direct injection

H₂ICE – hydrogen fueled internal combustion engine

H2O – water

HCLD - highly sensitive heated chemiluminescent gas analyzer

HEPA – high efficiency particulate arresting

HFID - heated flame ionization detector

IC – internal combustion

ISP – EEPS instrument specific parameter

MAP – manifold absolute pressure

MBT – maximum brake torque

NDIR – non-dispersive infrared detector

NH3 – ammonia

NI – National Instruments

NO – nitric oxide

NO₂ – nitrogen dioxide

NOx – oxides of nitrogen

xv

PCV – positive crankcase ventilation

PFI – port fuel injected

PID – proportional integral derivative

PM – particulate matter

RPM – revolutions per minute

SCR – selective catalytic reduction

SI – spark ignited

SO2 – sulfur dioxide

SO4 – sulfate fraction (in context of particulate origin)

SOF – soluble organic fraction of particulate

SOL – insoluble carbonaceous fraction of particulate

TDC – top dead center

THC – total hydrocarbons

TPM – total particulate mass

TWC – three way catalyst

UHC – unburned hydrocarbons

WOT – wide open throttle

1

1. Introduction

The reduction of exhaust emissions has been the focal point of engine development in the

automobile industry for the last several decades. The industry has taken many significant steps in

the last 50 years to reduce vehicle tailpipe emissions. Positive crankcase ventilation; two, and

then three-way catalytic converters; and exhaust gas recirculation (EGR) are just some of the

noteworthy advances that have been made. However, as vehicle emissions regulations get

stricter, the industry searches for near zero emissions solutions. There are many potential

solutions to this, each with their own advantages and disadvantages, but three of the most

popular are electric vehicles, hydrogen fuel cells, and hydrogen fueled internal combustion

engines.

Electric vehicles have received a considerable amount of attention in the last couple of

years. Electric vehicles are not new; in fact electric cars have been around since the late 1800’s

(Energy.gov, 2014). However, recent advances in batteries and motors have brought them to the

forefront. In general, the advantages of electric cars are that electricity is extremely inexpensive

and the distribution grid is already fully developed. The disadvantages are that the range of

electric cars is shorter than their gasoline counterparts, the recharging time is lengthy, and the

upfront cost is more than traditional gasoline fueled vehicles. Moreover, the environmental

impact of electric cars greatly depends on the method used to generate the electricity. Some

regions like Ontario generate most of their electricity with nuclear, hydro, or other renewables,

which makes the electricity fairly environmentally friendly. However, other jurisdictions produce

most of their electricity by burning fossil fuels, which in large part negates the benefit of electric

cars.

In general, electric cars are likely a good alternative to gasoline cars for short distance

daily commuting where the disadvantages of range and recharging time are less important.

However, consumers seem to be hesitant to purchase a vehicle that is more expensive and has

little personal benefit. Hybrid car sales have shown this trend. Hybrid car ownership has been

extremely flat, hovering at ~2.5% for the last decade (Nordan, 2013) (Cobb, 2015).

Fuel cell technology has been a hot topic for several decades but has consistently failed to

reach significant market penetration. The advantages of fuel cells are that they can have an

2

equivalent filling time to gasoline fueled vehicles and depending on the design, have

approximately the same range (The Washington Times, 2009) (Woody, 2014). The

disadvantages of fuel cells are that they are significantly more expensive than gasoline fueled

vehicles and they require very pure hydrogen to avoid poisoning the catalyst. Similarly to

electric vehicles, the cleanliness of hydrogen fueled vehicles depends on the source of the

hydrogen. Hydrogen can be processed in a very environmentally friendly way if it is electrolyzed

using renewable energy. However, most hydrogen is processed by reforming methane which has

negative environmental impacts.

Hydrogen fueled internal combustion engines are another possible replacement to

traditional gasoline fueled engines. Although they have gotten less publicity, there are several

advantages of hydrogen fueled internal combustion engines over the other two solutions.

Hydrogen fueled internal combustion engines are structurally very similar to traditional gasoline

fueled engines, so the costs are fairly similar. The only significant difference between gasoline

fueled engines and hydrogen fueled engines is the fuel storage and delivery system which

represents a fairly modest increase in cost. They are also less sensitive to the quality of the

hydrogen. The fuel is burned instead of reacted on a catalyst, so the purity is much less

important, which reduces the cost of the fuel. The disadvantage of hydrogen fueled internal

combustion engines is that they produce NOx which is a regulated emission, and because of

lubricating oil consumption, they produce particulate matter, carbon monoxide, and carbon

dioxide in small quantities.

A hydrogen fueled internal combustion engine with good oil control operating in the

correct conditions, would represent a very strong low emission replacement option for the

traditional gasoline fueled engine. The purpose of these experiments is to show that a well-

designed hydrogen fueled engine can produce ultra-low unburned hydrocarbon (UHC), NOx,

carbon monoxide (CO), carbon dioxide (CO2), and particulate matter (PM) emissions.

3

Origin of Test Engine 1.1.

Throughout the early 2000’s, a division at Ford worked on hydrogen fueled internal

combustion engines. Their objective was to develop a low cost replacement option for traditional

gasoline fueled engines that had ultra-low exhaust emissions. Over the course of about a decade,

the Ford team built several engines and published numerous journal papers on their progress.

The Ford team was working on a supercharged 2.3 L engine from a Ford Ranger when

the project was canceled. During the financial crisis in 2008, Ford went through a contraction

which included the termination of the hydrogen fueled internal combustion engine division. Ford

packed up the research engines that they were working on and donated them to several

universities. The University of Toronto was the recipient of one of these engines; a supercharged

2.3 L Ford Ranger engine. This engine was installed at the University of Toronto and used for

these tests.

4

2. Background

Hydrogen fueled engines are fairly unknown to most audiences, so a summary of the pertinent

topics will be presented. First, the unusual physical and chemical properties of diatomic

hydrogen will be discussed. Next, the governing principles and operating strategies of spark

ignited hydrogen fueled internal combustion engines will be explained. Many of the emissions of

hydrogen fueled engines are derived from the lubricating oil, so lubricating oil consumption

mechanisms and characteristics will be presented. Finally, particulate matter emissions and their

effect on human health will be examined.

Unique Properties of Hydrogen 2.1.

The public’s exposure to fuels are typically limited to gasoline, diesel, propane, and

natural gas. These four fuels have very different chemical and physical properties to hydrogen,

so a brief description of the differences is instructive. There are several interesting properties of

hydrogen which distinguish it from other fuels:

1. Hydrogen’s mass diffusivity into air is one of the highest of any known substance (Ng &

Lee, 2008);

2. It is the smallest known molecule (Segal, Wallace, & Keffer, 1986);

3. It has a very wide flammability limit (4 to 75%) (Segal, Wallace, & Keffer, 1986);

4. The minimum ignition energy is very low (0.02 mJ) (Segal, Wallace, & Keffer, 1986);

5. It has a high autoignition temperature (858 K) (Verhelst & Wallner, 2009);

6. The laminar flame speed is much faster than other fuels (290 cm/s) (Verhelst & Wallner,

2009);

7. Like many other flammable gases, hydrogen is colourless, odourless, and tasteless (Segal,

Wallace, & Keffer, 1986); and

8. A hydrogen flame is almost invisible (Segal, Wallace, & Keffer, 1986).

5

Hydrogen Fueled Internal Combustion Engines 2.2.

The following is a brief summary of Hydrogen Fueled Internal Combustion Engine

(H₂ICE) operational topics. One of the largest topics of H2ICEs are their propensity for abnormal

combustion. The various types of abnormal combustion events will be discussed followed by the

resulting impact on engine design. Next, the possible operating strategies of an H2ICE will be

presented and the effects on emissions will be shown. Exhaust Gas Recirculation (EGR) can play

an important role of H2ICE operation, so its effects on various engine systems will be explained.

Finally, the performance characteristics of H2ICEs will be discussed and compared to their

gasoline fueled counterparts.

2.2.1. Abnormal Combustion

At first glance, it would appear that a traditional port fuel injected (PFI) gasoline fueled

spark ignition (SI) engine could be easily retrofitted to run on hydrogen. The gasoline port fuel

injectors could be replaced with injectors meant for gaseous injection and with tuning of spark

timing and injection duration, the engine would be ready to go. However, this has proved to be

very far from the truth. The most serious difference between the use of hydrogen and gasoline in

an SI engine is the propensity for abnormal combustion. There are three types of abnormal

combustion referred to in the research literature: 1) surface ignition and pre-ignition, 2) backfire

or backflash, and 3) knock.

Surface ignition is the combustion of the in-cylinder hydrogen/air mixture on a hot

surface before the sparkplug fires. When the intake valve is closed, this ignition of the

hydrogen/air mixture is referred to as pre-ignition. Hydrogen/air mixtures have been shown to

ignite more easily than gasoline/air mixtures on hot surfaces (Swain, Swain, & Adt, 1988). Pre-

ignition can cause serious damage inside the combustion chamber if severe enough. This type of

abnormal combustion is fairly common and is mostly a function of the equivalence ratio, charge

density, spark timing, engine speed, and compression ratio (Tang, Kabat, Natkin, &

Stochhausen, 2002).

6

Backfire, sometimes referred to as backflash, is the propagation of a hydrogen flame

backwards through the intake valve and into the intake manifold. Backfire is a consequence of a

surface ignition event while the intake valve is still open and there is a flammable mixture that

leads out of the intake valve and into the intake manifold. Backfire is exacerbated with

increasing equivalence ratio. There are two reasons for this trend. As the equivalence ratio

approaches stoichiometry, the minimum ignition energy decreases, which tends to cause surface

ignition and thus backfire (Ciatti, Wallner, Ng, Stockhausen, & Boyer, 2006). Another less

important effect is that the injection event has to start earlier in the cycle. This is necessary to

produce a higher equivalence ratio. When the fuel is injected earlier in the cycle, the cylinder is

hotter because less time has elapsed since the last power stroke, so it is more likely to ignite on

the surface and propagate up the intake valve causing backfire (Ciatti, Wallner, Ng, Stockhausen,

& Boyer, 2006). These factors together mean that, as the equivalence ratio increases, surface

ignition and thus backfire become more prominent. Moreover, by retarding the spark timing, the

ignition event occurs later in the cycle, so the exhaust temperatures are increased. This also

causes the cylinder temperatures to increase which promotes surface ignition and backfire

(Ciatti, Wallner, Ng, Stockhausen, & Boyer, 2006). Backfire can pose a safety hazard if

flashback arrestors are not incorporated into the fuel system.

Although backfire is frequently referred to in past literature, it is now less common

because of better optimized combustion chamber design for hydrogen which reduces hot spots.

Some research teams have attempted to create a backflash event by altering the valve closing

times, injection timing, and injection flow rates in their research engines and were unable to

create a backflash event (Tang, Kabat, Natkin, & Stochhausen, 2002).

Knock is the most common form of abnormal combustion. The use of this term often

causes confusion because it is already used to define a similar event in a gasoline fueled SI

engine, but for a very different reason.

The occurrence of knock in a gasoline fueled SI engine is decided by a race between the

speed of the flame front and the auto ignition time of the end gases. In a gasoline fueled SI

engine, a spark fires in roughly the middle of the combustion chamber and a flame front passes

through the premixed fuel/air mixture. As the flame front passes through the mixture, the volume

of the burnt gas increases because of the increase in temperature which pressurizes the unburned

7

mixture. As the end gas, which is a mixture of unburnt fuel and air, is compressed, its

temperature increases. If the flame front fails to reach the end gas before it reaches its auto

ignition temperature, the end gas ignites volumetrically. This volumetric ignition of the end gas

can be very destructive to in-cylinder components and is one of the key design constraints to

many maximum operating conditions in gasoline fueled SI engines. In a gasoline fueled SI

engine, this problem is solved by increasing turbulence, reducing flame travel path (especially

over the hot exhaust valve(s)), using a higher octane fuel, and increasing the flame front’s

surface area as it proceeds through the combustion chamber.

In a hydrogen fueled SI engine, the term knock denotes a similar event with a very

different cause. In a hydrogen fueled SI engine, knock is caused when the subsonic flame front

that passes through the combustion chamber transitions to supersonic speeds because of

turbulence in the combustion chamber (Swain, Swain, & Adt, 1988). The supersonic speed of the

flame front results in a detonation of the end gases which cause a rapid pressure spike. This rapid

increase in pressure can seriously damage in-cylinder components. Research has shown that

advanced spark timing, increased engine speeds, and equivalence ratios closer to one increase the

propensity of the engine to knock (Ciatti, Wallner, Ng, Stockhausen, & Boyer, 2006). Given the

stark differences between the causes of knock in gasoline and hydrogen fueled SI engines, it

should be no surprise that the method for solving them is different. In a hydrogen fueled SI

engine turbulence should be minimized to ensure that the flame front does not transition to

supersonic speeds. This results in a very different cylinder and piston design for hydrogen

compared to traditional gasoline fueled SI engines.

2.2.2. Design Considerations

As discussed in the preceding section, the combustion characteristics of gasoline and

hydrogen are very different. So, the optimal engine design for a hydrogen fueled engine is very

different than for a gasoline fueled engine. The following list is a brief summary of typical

engine modifications that are made to gasoline fueled SI engines to make them run more

effectively on hydrogen:

8

1. One modification that can be made unilaterally to hydrogen fueled engines is a redesign

of the intake manifold shape. In a port fuel injected gasoline engine, the intake manifold

has narrow openings to increase inlet air velocity and ensure that the fuel is atomized at

low engine speeds. In a hydrogen fueled engine, atomization is not necessary because the

fuel is already a gas, so a wider intake manifold can be used which has a lower pressure

drop (White, Steeper, & Lutz, 2006)

2. The use of water cooled sparkplugs or low temperature sparkplugs which have less

thermal mass than traditional spark plugs are better because they have lower sustained

temperatures. The lower temperature sparkplugs reduce the probability of surface ignition

which can lead to knock and backfire (White, Steeper, & Lutz, 2006) (Swain, Swain, &

Adt, 1988).

3. By redesigning the combustion chamber to reduce turbulence, the flame speed of the

mixture will remain slow and help prevent knock (Swain, Swain, & Adt, 1988).

4. In traditional engine blocks optimized for gasoline, the coolant pathway designs can

result in slowed coolant flow in crevices which can cause film boiling. Film boiling

increases the temperature of the combustion chamber, especially the exhaust valve, which

can act as a source of surface ignition. By redesigning the coolant passageways in the

engine so that fluid flow is uniform and film boiling does not occur, a significant source

of surface ignition is removed (Swain, Swain, & Adt, 1988).

5. By using two exhaust valves instead of one, the heat transfer rate out of the exhaust

valves is doubled because the valve stem to guide contact area is doubled. This leads to

lower exhaust valve temperatures which decreases the chance of surface ignition and

knock (Swain, Swain, & Adt, 1988).

6. Hollow valves filled with sodium increase heat transfer rates out of the valve which

decreases the valve’s temperature. The exhaust valve in particular can serve as a source

of surface ignition in a hydrogen fueled engine, so sodium cooled exhaust valves have

successfully been implemented into hydrogen fueled research engines (Swain, Swain, &

Adt, 1988).

9

2.2.3. Operating Strategies

In a traditional gasoline fueled SI engine, the load of the engine is controlled with a

throttle. By opening the throttle, which is a valve in the air intake system, the amount of air

flowing into the engine is increased. The fuel/air ratio is kept relatively constant, and by

metering the mass of air flowing into the engine’s cylinders, the power output can be controlled.

The drawback of this operating technique is that the engine is using part of its power to pump air

across the pressure difference created by the throttle’s flow restriction. The piston moves

downward in the cylinder during the intake stroke and the pressure difference produced by the

partially closed throttle body results in an increase in work at the crankshaft.

In a compression ignition (CI) engine, often referred to as a diesel engine, the load of the

engine is controlled by metering the fuel injected. In a CI engine, only air is inducted into the

cylinder during the intake stroke and it is compressed to very high temperatures and pressures.

Fuel with a low auto ignition temperature is injected directly into the cylinder and the fuel and

air mix and burn spontaneously without the need for a spark. This system controls the power

produced by the engine by increasing or decreasing the mass of fuel injected directly into the

cylinder. In this control strategy, the engine takes in a constant volume of air per cycle regardless

of engine load. This results in reduced pumping losses and increased efficiency.

Much of the reason that gasoline fueled SI engines cannot successfully control load by

metering the fuel is tied to flammability limits. For a gasoline fueled SI engine, the equivalence

ratio must be relatively close to one for a stable combustion event with fast enough flame

velocities. However, hydrogen has a much faster flame velocity and wider flammability limits.

This means that the load of a hydrogen fueled SI engine can be controlled by metering the fuel

for much of its operating range (Tang, Kabat, Natkin, & Stochhausen, 2002).

Depending on specific engine characteristics, the highest equivalence ratio that a

hydrogen fueled port fuel injected SI engine without exhaust gas recirculation (EGR) can operate

at before knocking is approximately 0.7 (Natkin, et al., 2003). The lowest equivalence ratio that

it can operate at before there is significant deterioration in the combustion event is 0.1 (Verhelst

& Wallner, 2009). However, in-cylinder temperatures are very high when the equivalence ratio is

above 0.5, so NOx emissions are very high (Natkin, et al., 2003). Therefore, operating with an

10

equivalence ratio above 0.5 is rarely done. When the load is low enough, throttling of the intake

air is required while using an equivalence ratio of ~0.2. However, as the load increases, the

throttle body is opened until it is at wide open throttle (WOT). After this point, the equivalence

ratio can be increased to increase the load (Tang, Kabat, Natkin, & Stochhausen, 2002).

If EGR is used, the knock limited equivalence ratio is significantly increased because the

flame speed is reduced. Additionally, NOx levels are reduced with the use of EGR. However,

power is significantly reduced with the use of EGR. Therefore, the use of a turbocharger or

supercharger is required to generate more power. However, increasing the inlet pressure does

increase the cylinder temperature which increases the likelihood of knock. In practice, some

researchers have reported that the knock limited equivalence ratio goes from 1 to 0.5 when the

inlet pressure is increased from 1 to 2.6 bar absolute (White, Steeper, & Lutz, 2006). Intercooling

the boosted inlet air or injecting water into the combustion chamber, both of which reduce in-

cylinder temperatures, is typically required to increase the knock-limited equivalence ratio to

one.

The most fuel efficient strategy that still provides ultralow emissions changes depending

on load. When the engine is at idle or near idle conditions, the most efficient control strategy is

to operate the engine under fuel lean conditions while throttling the intake air (Verhelst &

Wallner, 2009). In this operating range there is no need for exhaust aftertreatment because the

NOx levels in the exhaust are already well below allowable limits. Moreover, at lean equivalence

ratios a Three Way Catalyst (TWC) would not work. For TWCs to operate, the exhaust must

oscillate between slightly rich and slightly lean. This reduces the NOx to N2 and O2 and oxidizes

the CO and UHCs to CO2.

For low loads, the throttle body should be set wide open and the engine’s power should

be controlled by metering the hydrogen flow rate (Verhelst & Wallner, 2009). Again, this

operating range creates very low levels of NOx, so no aftertreatment is necessary to meet

emissions guidelines. This control strategy can be used until the equivalence ratio reaches ~0.5

(Verhelst & Wallner, 2009). Increasing the equivalence ratio past this limit produces

unacceptably high NOx levels. At very low equivalence ratio, ~0.1, the combustion process

starts to deteriorate and unburned hydrogen levels in the exhaust increase. However, at

equivalence ratios above ~0.2 the combustion process is very stable and has a long combustion

11

duration in crank angle degrees. Although this is bad for fuel efficiency and power output, it

makes the engine idle much more smoothly. Moreover, all emissions are extremely low in this

operating range because in-cylinder temperatures are relatively low. All of these aspects put

together mean that hydrogen engines are much better overall at idle than gasoline fueled engines

(Ji & Wang, 2013).

During intermediate loads the engine can be controlled using the same strategy of Wide

Open Throttle (WOT) and fuel metering. To reduce the NOx emissions in the exhaust, a lean

NOx aftertreatment device such as Selective Catalytic Reduction (SCR) can be used.

Alternatively, the engine can be controlled with WOT and fuel metering below ~0.5 but with the

addition of a supercharger or turbocharger to increase power output (Verhelst & Wallner, 2009).

There are advantages and disadvantages to both SCR and boosting, so the decision between the

two should be made on a case by case basis.

At higher loads there is no well-established best practice; it depends heavily on the

system (Verhelst & Wallner, 2009). In general, the best control strategy is to use a fixed

equivalence ratio and control the engine load with throttling while simultaneously supercharging

or turbocharging and intercooling the inlet air (Verhelst & Wallner, 2009). In this mode the NOx

emission levels are low enough that exhaust gas after treatment is not necessary. Conversely, the

engine can be operated at stoichiometric conditions with a supercharger or turbocharger, Exhaust

Gas Recirculation (EGR) and a Three Way Catalyst (TWC) and the engine load can be

controlled using the throttle (Verhelst & Wallner, 2009) (White, Steeper, & Lutz, 2006).

2.2.4. Emissions

There is no carbon in hydrogen fuel, so the PM, UHC, CO, and CO2 emissions are solely

the result of combustion of the lubricating oil. Lubricating oil consumption rates are typically

fairly low, so the PM, UHC, CO, and CO2 emissions are usually very low. Therefore, the only

emissions which are normally of concern are NOx , H2, and hydrogen peroxide.

Advanced spark timing generally increases the NOx levels, but it has less of an effect

than the equivalence ratio (Ciatti, Wallner, Ng, Stockhausen, & Boyer, 2006). For most engine

12

setups, NOx levels are negligible when the equivalence ratio is below 0.45, but over 0.45 the

NOx emissions increase sharply (Ciatti, Wallner, Ng, Stockhausen, & Boyer, 2006). NOx levels

are significantly reduced with the use of EGR.

Much of the past research on hydrogen fueled engines was performed with the

assumption that the hydrogen was being burned completely even at very low equivalence ratios.

However, research has come to light to suggest that at very low equivalence ratios, a significant

portion of hydrogen is left unburned in the combustion chamber (Sinclair & Wallace, 1984).

Moreover, it has been shown that hydrogen peroxide can also be formed in hydrogen fueled

engines at very low equivalence ratios (Sinclair & Wallace, 1984).

The key reason for this unburned hydrogen in the exhaust at equivalence ratios below 0.5

is the reduction it causes in flame speed. When the equivalence ratio is below 0.5, the flame front

moves too slowly across the cylinder to consume the entirety of the hydrogen/air mixture

(Sinclair & Wallace, 1984). Although this explanation does not seem intuitive at first, it can be

presented in a more convincing manner. In most people’s personal experience; the combustion of

fuel/air mixtures seems to happen almost instantaneously. Although the reaction rates of

hydrogen/air mixtures are extremely fast, even at low equivalence ratios, the speed of the engine

is also very fast. For instance, if the engine is operating at 1800 rpm, which is a very common

operating condition for a vehicle’s engine, and each stroke is assumed to take up 180°, the power

stroke is only 17 milliseconds long. Now with the understanding that the timescales under

discussion are on the order of tens of milliseconds, it is believable that the flame front would not

have sufficient time to pass through the entirety of the hydrogen/air mixture at equivalence ratios

below 0.5.

Past research has indicated that hydrogen peroxide emissions are quite high when the

equivalence ratio is below 0.4. One study reported that at an equivalence ratio of 0.20, there was

1050 ppm of hydrogen peroxide in the exhaust stream (Adt, Swain, & Pappas, 1980). Although

this level is extremely high, subsequent research has been unable to replicate these results. More

typical hydrogen peroxide levels are ~150 ppm at very low equivalence ratios (Sinclair &

Wallace, 1984). Moreover, after a long length of pipe, the hydrogen peroxide levels were further

reduced to negligible levels. As a result, hydrogen peroxide emissions from hydrogen fueled

internal combustion engines serve little concern (Sinclair & Wallace, 1984). As long as the

13

exhaust system is sufficiently long with a relatively large internal surface area, the hydrogen

peroxide emissions in the exhaust will be negligible.

2.2.5. Exhaust Gas Recirculation

As was previously discussed, the knock limited equivalence ratio of a naturally aspirated

hydrogen fueled spark ignition engine is typically ~0.7 (Natkin, et al., 2003). However, at this

operating condition, a significant amount of NOx is formed. There are two solutions to this: 1)

run the engine leaner, or 2) use EGR and run the engine at an equivalence ratio of one. The

consequence of operating the engine at stoichiometric conditions is that a significant level of

NOx is produced. However, the high NOx levels can be taken care of with a Three-Way-Catalyst

(TWC) which typically reduces the NOx levels below five ppm (Natkin, et al., 2003). Overall,

the engine operating at an equivalence ratio of one with EGR and a TWC will have lower NOx

emissions than the engine operating at an equivalence ratio of ~0.5 with no exhaust

aftertreatment devices. However, the engine with EGR will produce less torque than the engine

operating at an equivalence ratio of ~0.5 (Natkin, et al., 2003). Therefore, a compromise must be

made between ultra-low emissions but slightly reduced torque, or fairly low emissions and

higher torque.

For gasoline fueled engines, the EGR rate is not directly measured because the required

measuring device is large and expensive. Instead, other sensors are used to infer the amount of

EGR being implemented based on CO2 levels in the exhaust stream. This causes a problem for

measuring the EGR rate in a hydrogen fueled engine because there are only trace amounts of

CO2 emissions from burning the lubricating oil, since hydrogen combustion produces no CO2

emissions. As a result, a best practice for measuring the EGR rate in a hydrogen fueled engine is

widely unknown. There are two accepted methods for measuring the EGR rate in a hydrogen

fueled engine: 1) constant volume method, and 2) O2 sensor method (Verhelst, et al., 2013).

The first method works on the principle that there is a fixed volume in the cylinder. If the

amount of air entering the cylinder is known when the EGR rate is zero, then the EGR rate can

be calculated if the instantaneous volumetric air flow rate, fuel flow rate, EGR pressure and EGR

temperature are known. This method has the benefit of simplicity, as it requires sensors that are

14

already commonly on engines, but it fails to take into account the effect of EGR on the

volumetric efficiency (Verhelst, et al., 2013).

With two wideband O2 sensors, one in the intake and one in the exhaust, the EGR rate

can be inferred based on a stoichiometric balance. Many engines already have one wideband O2

sensors in the exhaust for assessing the equivalence ratio at non-stoichiometric conditions, so

adding a second O2 sensor is relatively trivial. The only sensors it requires are two wideband O2

sensors, fuel flow rate and air flow rate. The advantage of this method is that it makes no

assumptions about engine operating conditions, for instance the volumetric efficiency can change

and the theory of the calculation still holds (Verhelst, et al., 2013).

For large production runs of vehicles, the first method is likely to be chosen because the

change in volumetric efficiency as the EGR rate increases can be found experimentally and then

programmed into all of the vehicles’ electronic control units (ECU). However, for laboratory

testing, the method of adding a second O2 sensor is probably more effective because it requires

no further assumptions.

2.2.6. Performance

Hydrogen fueled SI engines have lower torque and power output than their gasoline

fueled counterparts for several reasons. For port fuel injected SI engines, the fuel enters the

cylinder with the air. This inherently reduces the volumetric efficiency of the engine because less

air is inducted into the cylinder than is possible. In a port fuel injected gasoline fueled engine,

this effect does not significantly reduce the amount of air that is inducted into the cylinder

because the volume of gasoline required to operate the engine is relatively small. In a hydrogen

fueled port fuel injected SI engine however, the volume fraction of hydrogen is fairly significant.

For example, at an equivalence ratio of 0.5, the percentage of cylinder volume that the hydrogen

takes up can be determined by:

𝐻2 +0.5

∅(𝑂2 + 3.773𝑁2) → 𝐻2𝑂 + (

1.8865

∅)𝑁2 + (

0.5

∅− 0.5)𝑂2

15

𝜒𝐻2 =1

1 +(4.773)(0.5)

∅

=∅

3.3865= 0.1477

Therefore, at an equivalence ratio of 0.5, the hydrogen occupies ~15% of the volume that

is inducted into the cylinder. If this calculation is redone with gasoline, by modeling gasoline as

𝐶8𝐻18, the volume fraction is only 1.7% at an equivalence ratio of one. It is evident then that

even if everything else is the same, the power output of the hydrogen fueled engine will be

significantly lower.

The other significant factor in a hydrogen fueled engine’s reduced torque and power

output is its knock limited equivalence ratio. When the engine is operated with no EGR, the

knock limited equivalence ratio is ~0.7. As was mentioned earlier, in practice, the equivalence

ratio is rarely increased above ~0.5 because of NOx emissions. If EGR is used to increase the

equivalence ratio to one so that a TWC can be used, the torque and power output is further

reduced.

As a result, hydrogen fueled engines need to be supercharged or turbocharged to create

similar torque and power outputs as gasoline fueled engines. Unfortunately, pressurizing the

intake air also increases the intake air temperature which promotes knocking (Natkin, et al.,

2003). By adding a supercharger and intercooling the intake air, the torque and power output are

increased, but still not enough to be comparable to gasoline fueled engines (Natkin, et al., 2003).

However, if an A/C-to-air intercooler is added which further decreases the temperature of

the intake air, the hydrogen fueled engine produces equivalent torque and power output to a

gasoline fueled engine even with the parasitic losses of the A/C-to-air intercooler. (Natkin, et al.,

2003). This experiment shows that a hydrogen fueled engine can be operated with equivalent

torque and power output to a gasoline fueled engine given the correct control strategy and

modifications. This may seem like an unfair comparison because the gasoline engine would also

have a much higher torque and power output if it were supercharged or turbocharged and double

intercooled. Although this is true, it is a matter of reduced emissions. The gasoline fueled engine

would have much worse emissions than the hydrogen fueled engine even with the best emissions

16

technology. Therefore, it is a matter of keeping the performance that we are all accustom to from

a gasoline fueled engine while significantly improving the emissions. This all comes down to

cost. It is possible to make a hydrogen fueled engine with the same performance of a gasoline

fueled engine and significantly better emissions, but it will be more expensive than a regular

gasoline fueled engine.

Oil Consumption 2.3.

There are several sources of oil consumption in an SI engine: piston-ring-liner system,

valve stem seals, turbocharger, and crankcase ventilation (Froelund & Yilmaz, 2004). These

different sources are not all equal; they contribute different amounts depending on the load and

speed of the engine.

Figure 1. Oil Consumption Mechanisms (Froelund & Yilmaz, 2004)

There are four mechanisms that make-up the lubricant oil contribution from the piston-

ring-liner system (Froelund & Yilmaz, 2004):

1) Throw-off,

17

2) Transport through the piston ring groove,

3) Transport through the piston ring gap, and

4) Evaporation from the cylinder wall.

Figure 2. Lubricating Oil Consumption Mechanisms from the Piston-Ring-Liner System

(Froelund & Yilmaz, 2004)

Throw-off is the transport of liquid lubricating oil into the combustion chamber from

inertial forces of the piston ring assembly. The transport through the piston ring groove and gap

is the result of over pressurization of the piston’s second land region (Kapetanovic S. , 2009).

Research has shown that the gas flow through the top ring groove is much higher than through

the top ring gap (Froelund & Yilmaz, 2004). Contradictory to our traditional understanding of

lubricating oil desorption, lubricating oil is in fact desorbed into the combustion chamber during

all four strokes (Norris & Hochgreb, 1996). Moreover, it has been shown that lubricating oil is

oxidized during both the combustion event and the post flame stage (Norris & Hochgreb, 1996).

It has been shown in many studies that over the course of engine operation, the

constituents of the lubricating oil change considerably (Givens, et al., 2003). The lubricating oil

adsorbs engine fuel and retains heavy hydrocarbon and additives (Givens, et al., 2003). The

cause of this is thought to be the difference in volatility between different components of the

lubricating oil. As the engine heats up, lighter hydrocarbons in the lubricating oil evaporate and

18

escape the crankcase through various mechanisms. This leaves behind the heavier hydrocarbons

and organometallic compounds from the additives package.

There are several techniques available to measure the amount of lubricating oil being

consumed in the combustion chamber. One such technique, which has been used successfully for

decades, is the SO2 tracer method (Froelund K. , 1999). The working principle of the SO2 tracer

method is that the fuel being burned is of low, known sulfur content and the lubricating oil is of

higher, known sulfur content. With the air and fuel flow rate, and SO2 concentration in the

exhaust stream known, the lubricating oil consumption rate can be calculated. Another method is

to measure the change in volume or mass of the lubricating oil, but this method has many

disadvantages. If the volume or mass change method is used, there is no ability to measure the

effect of transient processes on lubricating oil consumption (Kapetanovic S. , 2009).

Additionally, hydrocarbon fuels adsorb into the lubricating oil which reduces the accuracy of the

method (Froelund K. , 1999). Finally, the test must be run for a very long time and the

lubricating oil must be completely drained for every test condition to measure the oil

consumption rate. This greatly increases testing time which reduces the number of operating

conditions that can be realistically tested.

However, in a hydrogen fueled SI engine there is no need to use the SO2 tracer method to

measure the lubricating oil consumption rate. In gasoline or diesel fueled engines, the SO2 tracer

method must be used because the measurement must be of a species which does not result from

the fuel. However, in a hydrogen fueled engine, there is no carbon in the fuel, so instead of

measuring the SO2 in the exhaust to determine the lubricating oil consumption rate, CO2 can be

measured instead. The advantage of measuring CO2 is that lubricating oil is mostly composed of

heavy hydrocarbons and a fairly small amount of it is sulfated ash. Therefore, the CO2 emissions

from the lubricating oil are much higher than the SO2 emissions. This increases the accuracy of

the lubricating oil consumption rate calculations which are based on the exhaust emissions

measurements. Additionally, measurement of CO2 is more common, so the exhaust measurement

equipment tends to be less expensive, more accurate, and more readily available.

It has been known for a long time that lubricating oil consumption increases as the load

and speed of the engine increases and that the lubricating oil consumption has a periodicity

(Froelund K. , 1999). By periodicity we mean to say that the oil consumption cyclically increases

19

and decreases with a steady frequency. This periodicity is thought to be the result of rotation of

the top piston ring.

Another trend sometimes observed in lubricating oil consumption rates is the spike

caused at intermediate loads and speeds. At a particular speed and load of some engines,

typically intermediate speeds and low loads, the oil consumption rate increases far above normal

levels (Froelund K. , 1999). If the operating conditions are altered slightly from this position, the

lubricating oil consumption rate decreases back to normal levels. It is thought that this unstable

operating condition is the result of roughly equal gas forces and inertial forces on the top piston

ring. These roughly equal forces on the piston ring cause the piston ring to chatter which reduces

sealing and increases lubricating oil consumption.

Engine Particulate Emissions 2.4.

Particulate matter formation occurs in engines when hydrocarbons are heated in the

absence of oxygen. If the PM is only composed of carbon, it is referred to as soot. Soot forms

through four steps: 1) pyrolysis, 2) nucleation, 3) coalescence, and 4) agglomeration (Tree &

Svensson, 2007). If there is no oxygen present when the fuel is heated, it forms soot precursors,

polycyclic aromatic hydrocarbons (PAHs) through pyrolysis. Simply put, pyrolysis is a process

where the molecular structure of the fuel reconfigures itself by removing hydrogen atoms (Tree

& Svensson, 2007). This removal of hydrogen increases the C/H ratio of the molecule. Again, if

no oxygen is present, nucleation of the gas phase PAHs into solid phase particles occur.

Coalescence is the continued surface growth of the particles formed during nucleation.

Agglomeration is the process where several particles, enlarged by the coalescence process, come

together to form complex branching structures. Although these processes are described in a

manner suggesting that they happen chronologically, in reality, all four processes take place

nearly simultaneously. At any point during the processes, oxidation of the particles can occur

which would result in CO or CO2 production.

However, in a real engine, the PM is not only composed of soot. The total particulate

mass (TPM) is typically divided into three categories; the 1) insoluble carbonaceous fraction

20

(SOL) which is typically referred to as soot, 2) soluble organic fraction (SOF), and 3) sulfate

fraction (SO4) (Sappok & Wong, 2007).

Figure 3. Current Understanding of the Structure of a Complex Engine Exhaust Particle (Matti

Maricq, 2007)

With particulate emission regulations becoming stricter, the method used to measure the

PM emissions is becoming more important. As PM emission regulations reduce allowable levels

below 5 mg/km, the temperature, electrostatic discharge, humidity, and gas phase adsorption

during the test results in measurement noise on the same scale as the actual PM being measured

(Matti Maricq, 2007).

Moreover, with our enhanced understanding of the formation of PM emissions and PM

emissions’ effect on human health, there is debate over the best way to quantify and regulate PM

21

emissions from vehicles. Historically, the mass of PM emissions per kilometer has been the

regulated parameter, but recent research suggests that small PM emissions are more harmful to

human cardio-pulmonary systems than large PM emissions (Burtscher, 2005). Additionally, the

chemical composition of the surface of PM emissions has been shown to have strong effects on

human health (Burtscher, 2005).

Another issue surrounding the regulation of PM emissions is the method used to quantify

them. The creation of PM emissions is not limited to the combustion chamber (Burtscher, 2005).

The process of cooling and diluting the exhaust stream has a strong effect on PM formation. The

optimal test setup would be one which replicates the conditions found in the real world where the

engine’s exhaust system discharges into the atmosphere. However, as one can imagine, this does

not result in a discrete set of conditions. The environment in which engines operate changes

drastically. The temperature, pressure and wind conditions are all aspects which directly affect

PM formation (Burtscher, 2005). Additionally, there is no clear dilution level and resonance time

that should be replicated to apply to real world human exposure. People are in contact with the

exhaust from engines at various distances, whether it be walking on the sidewalk next to running

cars or sitting in a park a great distance from the street. There is no obvious set of circumstances

to replicate in laboratory engine testing.

Research also shows that there are two distinct particle families: soot, which tends to be a

larger formation of solid particles; and volatile materials, which tend to be smaller (Burtscher,

2005). Soot forms through the accumulation of small carbon formations and volatiles form

through nucleation.

Nucleated volatile particulate formation is more sensitive to the parameters of the exhaust

process than soot particles. This means that depending on the exhaust system setup, volatile

particles can outnumber soot particles or may vanish altogether (Burtscher, 2005). In general,

catalytic aftertreatment devices aid in the nucleation process, helping to produce more volatile

particulate. In real world situations, the engine’s exhaust is rapidly diluted into the environment

which quickly decreases the temperature. This rapid cooling process strongly favours nucleation

where almost all of the volatiles transition to the particle phase (Burtscher, 2005).

22

2.4.1. Particulate Emissions from Lubricating Oil

In traditional gasoline fueled engine, particulate matter formation occurs from two

sources: 1) the fuel, and 2) the lubricating oil. Lubricants are composed of 70-83% refinery-

derived organic base stocks, 5-8% viscosity modifiers, and 12-18% inorganic additives (Sappok

& Wong, 2007).

Organometallic compounds from the additives package and hydrocarbons are evaporated

during engine operation and may be partially oxidized depending on conditions. Oxidized

metals can serve as nucleation sites for particle growth by carbon addition (Miller, Stipe, Habjan,

& Ahlstrand, 2007). Particles can grow by agglomeration or by particle adsorption. In PFI SI

engines, this is virtually the only source of PM emissions. However, in DI SI and CI engines, as

PM formation from the fuel is reduced, PM formation from the lubricating oil is becoming more

important. As a result, particulate matter emissions which result from the lubricating oil have

become a hot topic. There are three main reasons for this increased attention in lubricating oil

derived particulates.

In compression ignition (CI) engines (also known as Diesel engines), a portion of the PM

which originates from the lubricating oil, referred to as ash, cannot be regenerated out of a Diesel

Particulate Filter (DPF) (Czerwinski, Petermann, Ulrich, Mueller, & Wichser, 2005). The

inorganic additives are responsible for most of the ash emissions in a diesel engine (Sappok &

Wong, 2007). Figure 4 shows the cross section of a DPF after regeneration where ash remains on

the surface. So, over a long period of time, as the DPF fills with ash from the lubricating oil, the

engine’s power and efficiency decreases (Czerwinski, Petermann, Ulrich, Mueller, & Wichser,

2005).

Figure 4. Condition of Diesel Particulate Filter after Regeneration (Givens, et al., 2003)

23

Moreover, lubricating oil consumption is known to contribute significantly to the SO2

emissions in the exhaust. This is becoming especially important as the sulfur level in diesel fuel

steadily declines to adhere to new government regulations. It has been shown that SO2 adsorbs

onto NOx storage catalysts which reduces the effectiveness of the catalyst (Givens, et al., 2003).

The SO2 level in the storage catalyst can be significantly reduced after regeneration periods, but

regeneration should be delayed as long as possible to maintain fuel economy (Givens, et al.,

2003).

The other reason for interest in particulate matter formation as a result of lubricating oil is

the effect of lubricating oil derived PM on human health. It has been speculated that lubricating

oil derived PM is particularly bad for human health because of the organometallic compounds

which result from the additives package in lubricating oils.

As a result of these two driving forces, much more is now understood about the

consumption of oil in internal combustion engines and the impact on PM emissions. Lubricating

oil from the various transportation pathways described in Section 2.3 have different lifecycles in

the engine, so they contribute to particulate emissions in different ways. In general, the piston-

ring-liner system contributes the most to lubricant derived particulate emissions, especially at

higher speeds and loads (Froelund & Yilmaz, 2004).

In addition to furthering our understanding of lubricating oil consumption mechanisms,

the effect of lubricating oil properties on consumption rates has also been studied. In general, oil

viscosity and volatility are thought to be paramount in the determination of particulate emissions

(Froelund & Yilmaz, 2004). The viscosity of the oil affects the transport of liquid oil into the

combustion chamber and the volatility affects the vapourization of the liquid oil on hot surfaces.

It is evident then that increased viscosity and decreased volatility reduces lubricating oil

consumption (Froelund & Yilmaz, 2004). However, a balance needs to be struck between

lubricating oil consumption rates and other factors. As the viscosity of the lubricating oil is

increased, the friction in the engine increases which increases in-cylinder temperatures. This

increase in temperature results in higher NOx emissions (Andersson, Preston, Warrens, & Brett,

2004) (Froelund & Yilmaz, 2004).

24

It has been shown that most of the PM emissions which result from the lubricating oil are

in the nucleation range (Andersson, Preston, Warrens, & Brett, 2004). It has also been generally

shown that increasing the amount of sulfated ash or phosphorous in the lubricating oil increases

the PM emissions (Andersson, Preston, Warrens, & Brett, 2004). However, the oil consumption

and resulting PM emissions from an engine are extremely complex processes. A similar study

which assessed the PM emissions with different lubricating oils observed the opposite trend. By

increasing the sulfate level in the lubricating oil, the PM emissions decreased (Czerwinski,

Petermann, Ulrich, Mueller, & Wichser, 2005). This is a very unusual result which is far out-of-

line with similar research. But it does go to show that with such a complicated system, changing

even one parameter can have far-reaching effects on many systems.

A hydrogen fueled engine can be used to assess the impact of lubricating oil on

particulate matter emissions. One study that used a hydrogen fueled engine to measure the

effects of lubricating oil on PM emissions encountered five types of particles (Miller, Stipe,

Habjan, & Ahlstrand, 2007):

1. Agglomerates are defined as being between 100 and 400 nm in diameter and mainly

composed of carbon with a small amount of other elements. This type of particle is by far

the most common in diesel engines, but it is much less common in hydrogen fueled

engines. Therefore, it is most likely a result of diesel fuel.

2. Dense spheres, with diameters ranging between 30 and 300 nm and mainly composed of

metals, are common in hydrogen fueled engines. These particles are denser than

agglomerates, likely because of their high calcium content, and are thought to be the

result of lubricating oil constituents.

3. Less-dense spheres are of the same diameter range as dense spheres (30-300 nm), but

consist of less calcium and therefore have a lower density.

4. Core-shell particles are made up of a dense core and a less-dense coating. Lubricating oil

in the vapour phase condenses onto nucleated metallic spheres as the in-cylinder mixture

is cooled during the expansion stroke or as the exhaust gases exit through the exhaust

system.

25

5. Nanoparticles, which have diameters ranging from 5 to 50 nm and mainly consist of iron

and carbon, are also observed in the exhaust stream. These particles are thought to form

at high temperatures shortly after the combustion event.

Human Health Effects of Particulate Emissions 2.5.

It has been shown with numerous epidemiological studies that life expectancy is reduced

with increasing environmental PM emissions (Pope III, Ezzati, & Dockery, 2009) (Boldo, et al.,

2006). One such study showed with regression analysis that an average decrease in

environmental PM concentration of 6.52 μg/m³ over a roughly 20 year period across 51 U.S.

metropolitan resulted in an average increase of 0.4 years to life expectancy (Pope III, Ezzati, &

Dockery, 2009). Another study in Europe showed that a decrease in particulate matter

concentration to 15 μg/m³ across 23 cities would prevent 16926 premature deaths annually

(Boldo, et al., 2006).

Internal combustion engines have historically contributed significantly to environmental

PM emissions. However, there are many sources of PM emissions, both natural and

anthropogenic. Most of the epidemiological studies use large data sets which do not distinguish

environmental PM concentrations by source. Therefore, it should be noted that most of the

effects of PM on life expectancy are from total environmental PM emissions and not solely from

internal combustion engines. Internal combustion engines serve as an important part of the

environmental PM emissions, but only a part.

It has also been shown that organometallic derived exhaust particles have an

overwhelming effect on the lung tissue of mammals (Ghio, Richard, Carter, & Madden, 2000).

This organometallic derived PM is the result of the lubricating oil. This result shows that

lubricating oil consumption effects are potentially very important to PM emissions and resulting

human health impacts.

26

3. Experimental Setup

The engine was operated in both naturally aspirated and supercharged configurations for the

experiments. The exhaust sampling equipment was the same for the naturally aspirated and

supercharged tests.

Naturally Aspirated Engine Configuration 3.1.

Figure 5 shows the order of primary components including the emissions sampling

equipment for the naturally aspirated configuration. The emissions equipment will be described

in more detail in section 3.5.

Figure 5. Diagram of Naturally Aspirated Engine Configuration

Table 1 shows the general specifications of the engine used for the experiments. This

engine design was first intended for gasoline fueled operation. However, several alterations were

made by Ford for hydrogen operation.

27

Table 1. Engine Specifications

Manufacturer Ford

Intended Vehicle Ranger

Production Year 2001

Number of Cylinders 4

Cylinder Orientation In-line

Fuel Injection Style Port

Displacement 2.3 L

Compression Ratio 12.2:1

Number of Valves Per Cylinder 4

Camshaft Style Dual Overhead

The following upgrades were made by Ford to facilitate hydrogen operation (Natkin, et

al., 2003):

The traditional cast eutectic alloy piston with a 3.5 mm second ring land width was

replaced with a forged eutectic alloy piston with a 5.5 mm second ring landing

The original 21 mm pressed pin was replaced with a 23.1 mm floating piston pin

Several modifications were made to the connecting rods

o The top hole was fitted with a bronze pin bushing

o The stock connecting rod was replaced with a sturdier version that incorporated

an H-beam cross sectional shape

o The new connecting rod was 2.38 mm shorter than the stock version to facilitate

the other changes made

The valve seat inserts were made out of hardened tool steel (50-60 Rockwell C)

The valves were faced with a Stellite seat for better wear resistance

Finish honing was performed on the cylinder block to get an average cylindricity of 6-7

μm

Customized piston rings were used, but no specific details were provided by Ford

The valve stem seals were upgraded from a Grade 2 to a Grade 1 seal by applying a

diamond-like coating to the valve stems that contact the seals

28

Figure 6 shows a picture of the engine configuration for the naturally aspirated tests.

From this view of the engine, all of the important components can be seen.

Figure 6. Picture of Naturally Aspirated Engine Configuration

Supercharged Engine Configuration 3.2.

For the supercharged experiments, Figure 7 shows the layout of the primary components.

For the supercharged configuration, a plenum, supercharger, and intercooler were added to the

naturally aspirated configuration.

Figure 7. Diagram of Supercharged Engine Configuration

29

The supercharger used for the experiments is twin screw style (Part Number:

KJ05207AX) that is driven off of the Front End Accessory Drive (FEAD). This model of

supercharger was used in production on the 2001 Mazda Millenia.

There are three main types of supercharger:

Roots

Twin-screw

Centrifugal

Roots and twin-screw superchargers are positive displacement pumps. For both of these

supercharger types, two rotating sets of blades mesh with each other and squeeze the air to a

higher pressure at the exit. The main difference between Roots and twin-screw superchargers is

the shape of the blades that mesh with each other. Roots superchargers force air along the

periphery of the supercharger walls whereas twin screw superchargers force the air in between

the two sets of blades. The advantage of both of these types of superchargers is that they produce

a constant outlet pressure with speed. This means that they supply a significant quantity of air

into the engine at low speeds. This adds torque and power to the engine at low speeds. The major

disadvantage of this type of supercharger is that it consumes a lot of power which considerably

reduces the fuel conversion efficiency of the engine.

Centrifugal superchargers have a set of vanes on a rotor inside of a circular housing. As

the rotor spins faster, the pressure at the exit of the supercharger increases. The advantage of this

type of supercharger is that it consumes less power, which increases engine efficiency as well as

the final output power at maximum engine speed. The disadvantage is that a centrifugal

supercharger produces almost no pressure at low speeds and therefore does not increase the

engine’s power at low speeds.

When this engine was under development by Ford, they were attempting to recreate the

power of a gasoline fueled engine on hydrogen. For this reason, they needed the supercharger to

produce compressed air throughout the speed range of the engine. As such, a twin-screw

supercharger was selected.

30

Figure 8 shows the engine configuration for the supercharged tests. As the picture shows,

the intake piping (blue) now includes the plenum, supercharger, and intercooler. All of the

emissions sampling equipment was the same for the supercharged and naturally aspirated tests.

Figure 8. Picture of Supercharged Engine Configuration

Positive Crankcase Ventilation 3.3.

The engine crankcase has to be ventilated to prevent the buildup of flammable and

corrosive gases that escape past the piston rings into the crankcase (blow-by). Positive Crankcase

Ventilation (PCV) is used to recycle the gas (including unburned hydrocarbons) in the crankcase

to the engine’s intake to be burned rather than being released into the atmosphere. For gasoline

fueled engines, these unburned hydrocarbons in the crankcase come from two sources: 1)

unburned fuel from the combustion chamber that slips past the piston rings during the

combustion event, and 2) vapourized lubricating oil which is entrained into the gas from the

31

agitation of the oil in the oil pan. In a hydrogen fueled engine, the UHCs in the crankcase cavity

are solely from the lubricating oil.

A traditional PCV system routes the vapour space of the crankcase cavity to the intake

manifold. By routing the crankcase vapours to the intake manifold, the UHCs are burned in the

combustion chamber and turned into CO₂ which is much better for the environment. In a

supercharged engine, the PCV system layout has to be more complicated because the intake

manifold is sometimes above atmospheric pressure.

The PCV system for this test needed to be able to switch between a supercharged and

naturally aspirated configurations. The PCV layout for the naturally aspirated tests followed the

traditional layout as seen in Figure 9. The PCV system for the supercharged case was based on a

system Ford developed for their hydrogen fueled engines; see Figure 10.

Figure 9. Naturally Aspirated PCV System (G2IC Turbo Guide, 2016)

32

As Figure 9 shows, the crankcase vapour is routed to the intake manifold downstream of

the throttle plate. A PCV valve ensures that the flow direction does not reverse.

Figure 10. PCV System Diagram for Supercharged Configuration (Natkin, et al., 2003)

The PCV system for the supercharged configuration closely followed the layout

developed by Ford in Figure 10. As shown in Figure 10, the crankcase vapour is routed through

an oil coalescing filter before being sent to either the intake manifold or a Venturi at the inlet of

the supercharger. The crankcase vapours flow to the intake manifold when the intake manifold is

at negative pressure. When the boost pressure is being used by the engine and the intake

manifold is at positive pressure, the crankcase vapours flow to the inlet of the supercharger.

33

Although the PCV system for the experiments closely followed this flow diagram, there

are some notable differences. The drawing of the supercharger in Figure 10 indicates that it is a

centrifugal supercharger. However, the supercharger used for these experiments is a twin-screw

supercharger. Furthermore, the intercooling heat exchangers for these experiments are different

than the ones in Figure 10. In Figure 10 there are two intercoolers, an air-to-air intercooler and

an air-to-AC intercooler. In these experiments there is only one intercooler downstream of the

supercharger. With only one intercooler, there is no need for the bypass stream just right of the

supercharger inlet Venturi in Figure 10. An air-to-water intercooler is used for these experiments

and the domestic water supply is used as the coolant.

3.3.1. Oil Coalescing Filter

The oil coalescing filter used for the tests was a Mann+Hummel ProVentⓇ 200. This oil

coalescing filter was chosen because it has a very high filtering efficiency.

Figure 11. Oil Coalescing Filter (MANN+HUMMEL ProVent, 2016)

34

Switching Between Supercharged and Naturally Aspirated 3.4.

Configurations

The engine was operated both naturally aspirated and supercharged for the experiments.

The procedure to switch from the supercharged to naturally aspirated engine configuration, or

vice versa, was a fairly streamlined process. There were three items on the engine that needed to

be altered to switch operational mode.

The supercharger was powered from the Front End Accessory Drive (FEAD). A

serpentine belt connected the crank, water pump, belt tensioner and supercharger. Two different

sized serpentine belts were used for the tests, a larger one for the supercharged tests and another

smaller one for the naturally aspirated tests.

The Positive Crankcase Ventilation (PCV) system, which was described in section 3.3,

was another aspect that needed to be changed when switching operational mode. Quick

disconnects were used on the PCV system tubes to facilitate switching the operational mode of

the engine quickly.

Finally, the intake piping was changed for the two operational configurations. In the

supercharged configuration, a pipe connected the outlet of the intercooler with the inlet of the

throttle body. For the naturally aspirated configuration, this pipe was removed and replaced with

a mass airflow sensor which drew air from the test cell.

Exhaust Emissions Equipment 3.5.

There are six exhaust emissions sampling instruments used for these experiments. Two of

the instruments measure aspects of the particulate matter in the exhaust and the other four

measure gaseous concentrations of exhaust species.

35

3.5.1. Isokinetic Probe

It is common in engine testing experiments to draw a known flow rate of diluted engine

exhaust through a clean, weighed filter for a specified amount of time so that the particulate mass

as a function of time can be determined. However, for this engine, the particulate levels in the

exhaust are so low that a special probe and sampling setup are needed.

A customized isokinetic probe was designed and manufactured for these tests. The

purpose of the isokinetic probe is to draw a raw sample from the exhaust without affecting the

particles in the sample. The isokinetic probe sits inline in the exhaust stream and draws the

sample in at the same velocity as the exhaust stream passing next to the isokinetic probe. Figure

12 shows the conceptual design of the isokinetic probe used in the experiments.

Figure 12. Isokinetic Probe Diagram

In order to capture a representative PM sample, it is very important that the sample

flowrate is correct. If the sample flow rate is too high, the results will overestimate the mass of

36

particles because the number of particles being drawn into the isokinetic probe will be

disproportionately high. Moreover, too high a sample flow rate will skew the particle

distribution. Small particles are more influenced by pressure differences in the flow because they

have less inertia. As a result, small particles will be overrepresented in the filter sample. On the

other hand, if the isokinetic probe is drawing too low a flow rate, the particle mass will be too

low and the distribution will disproportionately favour large particles.

A pump is used to draw a sample through the weighed filter. To ensure that the sampling

rate is correct, a needle valve at the outlet of the pump is adjusted. Based on previous work on a

similar sampling setup, it is known that for the sampling velocity to match the exhaust stream

velocity, the static pressure of the exhaust must match the static pressure at the tip of the

isokinetic probe. A monometer is used to observe the difference between the two static pressures.

The needle valve at the outlet of the pump is adjusted until the manometer is balanced on both

sides. Figure 13 shows the actual isokinetic probe used in the tests.

Figure 13. Isokinetic Probe

37

In typical particulate sampling setups, the exhaust sample is diluted with filtered dilution

air before being routed through the weighed filter. However, in these experiments, because of the

low particulate levels in the exhaust, the filter samples are taken on an undiluted basis. This raw

exhaust sampling causes many complications.

To draw a sample out of the exhaust, a pump is used. However, the majority of pumps are

incapable of surviving at exhaust temperatures, so the sample needs to be cooled before being

put through the pump. All engine exhaust, but especially hydrogen fueled engine exhaust, has a

lot of water in it. So, when the engine exhaust is cooled, most of the water condenses out. Few

air pumps can withstand having water pumped through them, so the liquid water needs to be

removed. Finally, the flow rate needs to be determined. A rotameter is used to measure the

volumetric flow rate, but the mass flow rate is needed. A second manometer with one end

connected to the exhaust and the other end open to the room is used to determine the pressure at

the sampling point. Finally, a thermometer is used at the outlet of the rotameter to measure the

temperature. With the volumetric flow rate, pressure, and temperature; the mass flow rate going

through the filter can be determined.

Figure 14. Isokinetic Sampling Flowchart Diagram

38

As seen in Figure 14, a sample is drawn isokinetically into the probe. The sample flows

through the probe and ball valve and into the filter. After the exhaust sample flows through the

filter and the particulate matter present in the exhaust sample is deposited on the filter element,

the filtered sample passes through a heat exchanger. After being cooled with coolant water in the

heat exchanger, the condensed water drips into a water drain and the remaining sample exits the

heat exchanger. Although most of the water has been removed from the sample, the sample is

still very humid and any further cooling will result in more condensation forming. Therefore, the

sample is routed into a desiccant to further dry it before entering the pump. The rotameter is a

very sensitive device. If any dust or particles from the pump are put through it, the measuring

element will be clogged and will not work. Therefore, the sample is cleaned once more with a

filter before entering the rotameter. After flowing through the needle valve and rotameter, the

sample is returned to the exhaust stream.

Figure 14 also shows the two manometers on the right hand side. The left manometer

measures a pressure difference that is balanced (zero pressure differential) by adjusting the

needle valve on the rotameter. This is to ensure that the sample flow rate into the isokinetic probe

is correct. The right manometer is used to measure the pressure at the exit of the rotameter. This

pressure is needed to relate the volumetric flow rate of the rotameter to a mass flow rate.

3.5.2. Engine Exhaust Particle Sizer

In addition to the isokinetic probe setup which is used to gather gravimetric results, the

Engine Exhaust Particle Sizer (EEPS) is used to obtain size distribution and concentration

results. There are four components in this setup:

1. TSI MD19-3E Rotating Disk Diluter

2. TSI 379020A Rotating Disk Thermodiluter

3. TSI 379030 Thermal Conditioner Air Supply

4. TSI 3090 Engine Exhaust Particle Sizer

39

Figure 15. Pump (Left) and Diluter (Right) Configuration for EEPS (Matter Engineering, 2014)

Figure 15 shows the MD19-3E on the left connected to the 379020A diluter on the right.

The MD19-3E rotating disk diluter consists of a positive displacement pump with a ten cavity

disk and a first stage diluter. This component draws a sample out of the engine’s exhaust and

dilutes it with HEPA filtered air.

Next, the drawn, diluted sample is sent to the 379020A rotating disk thermodiluter. The

379020A serves two purposes, it has all of the controls for the rotating disk diluter and it serves

as another dilution stage.

The 379030 thermal conditioner air supply is the final stage of dilution. This component

dilutes the sample so that the flow rate is high enough for the EEPS. The 3090 EEPS requires a

10 L/min flow rate which the 379030 thermal conditioner air supply provides.

40

The 3090 EEPS is the particulate matter analyzer. It measures the concentration of

particulate matter in bins that are separated by particle diameter. This machine has been used in

previous experiments and a correction was established by former experimenters (Zimmermann,

et al., 2014). The EEPS, which is a dynamic PM measuring instrument, was compared to a very

sensitive steady state PM measuring instrument. The correction developed by (Zimmermann, et

al., 2014) was used for all of the data presented in this thesis.

Finally, the dilution ratio is needed to correct the PM concentrations to a raw exhaust

basis. The exhaust sample, which is drawn out of the exhaust, is diluted in several stages before

the PM concentration is measured by the EEPS. Therefore, the PM concentration measured by

the EEPS is on the diluted basis. To determine the PM concentration on a raw exhaust gas basis,

the dilution ratio is needed. The settings on the 379020A can be used to calculate the dilution

ratio.

In previous experiments, the dilution ratio was verified by comparing the CO₂

concentration measured at the outlet of the EEPS and in the exhaust. However, in these

experiments, the hydrogen fueled engine produces only trace amounts of CO₂. Therefore, the

dilution ratio cannot be determined by comparing the CO₂ concentration of the raw exhaust with

the CO₂ concentration of the diluted mixture. Instead, the settings on the 379020A are used to

calculate the dilution ratio. Although this is not as accurate a method, previous experiments on

hydrocarbon fuels have shown that the dilution ratio predicted by the settings on the 379020A is

reasonably close to the dilution ratio calculated with the CO₂ method.

Table 2. EEPS Dilution Settings

Primary Dilution Temperature (°C) 80

Primary Dilution Factor (%) 100

Secondary Dilution Factor (V) 6.5

Thermal Conditioner (°C) 300

Calculated Dilution Ratio 100

𝐷𝑅 =𝐼𝑆𝑃 × 𝐷𝑆 × 𝐶𝑇

𝐷𝑃=(1543)(6.5)(1.0)

100= 100.295

41

Where:

DR ~ Dilution ratio

ISP ~ Instrument specific parameter

DS ~ Secondary dilution factor

CT ~ Coefficient based on the primary dilution temperature

DP ~ Primary dilution factor

3.5.3. Fourier Transform Infrared Spectroscopy

A 2030HS Fourier Transform Infrared Spectroscopy (FTIR) from MKS was used to

measure various gaseous species. The FTIR works on the basis of exposing an exhaust sample to

a laser. The absorption of the laser in a sample is measured and compared to the absorption

pattern of known compounds. There are two large advantages of this exhaust analyzer:

The sample is kept at 191°C which means that the sample does not need to be dried. This

allows all of the measurements to be made on a wet basis. It also allows the FTIR to

measure the water concentration in the exhaust.

When the FTIR collects data, it is actually collecting absorption spectra data. This means

that the raw spectral data can be stored and rerun with different recipes, looking for

different species in the exhaust.

A ‘recipe’ is used for the FTIR to search for specific gaseous species in the sample. The

recipe takes the measured spectra from the test and compares it to spectra of known species at

known concentrations. Based on the absorption of the spectra and the wavelength of that

absorption, the FTIR uses Fourier transforms to compute the concentration of different species.

The recipe is needed because species have overlapping spectral interferences. The FTIR is

incapable of looking for all species simultaneously because there would be too many overlapping

interference regions.

42

For these experiments, the following species were included in the recipe:

H₂O

NO

NO₂

CO₂

CO

Formaldehyde

NH₃

3.5.4. Emissions Bench

The emissions bench from California Analytical Instruments (CAI) was used to measure

all of the gaseous emissions regulated by the EPA. The emissions bench measures THC, NOx,

CO₂, O₂, and CO. The oxygen measurement from this instrument was used for all of the

equivalence ratio calculations which are used extensively in the results section. Moreover, the

NOx measurements from this instrument were compared to the FTIR’s NOx measurements.

Since NOx is the only regulated emission produced by hydrogen engines in any significant

quantity, two measurements are beneficial to ensure agreement between instruments.

The other analyzers for THC, CO₂, and CO are less important for these specific

experiments. Although hydrogen fueled engines produce THC, CO₂, and CO from combustion or

incomplete combustion of the lubricating oil, the levels are so low that they are difficult to

measure with most instruments. The analyzers for measuring THC, CO₂, and CO were calibrated

and used for all of these tests, but their results are far too low to be statistically significant based

on the sensitivity of the analyzers. For CO, only the FTIR is sensitive enough to measure the

level produced by the engine. As for CO₂, both the FTIR and LICOR (described below in 3.5.5)

are sensitive enough to measure the concentration.

43

Table 3. Emissions Bench Channels and Ranges

Emissions Analyzers Calibration Cylinders

Model

Number Analyzer Type Species Ranges Concentration Species

CAI 600

HFID

Heated Flame

Ionizing Detection THC – C₃

Basis

300 ppm 203 ppm C₃H₈

3000 ppm 2000 ppm

CAI 600

HCLD

Heated

Chemiluminescence

Detection

NOx

100 ppm 89.7 ppm

NOx 1000 ppm 900 ppm

5000 ppm 4063 ppm

CAI 601P

NDIR

Non-Dispersive

Infrared and

Paramagnetic

CO₂ 9.0 % 9.0 %

CO₂ 14.0 % 13.5 %

O₂ 1.0 % 0.99 % O₂

21 % 20.946 % O₂

CAI 602

NDIR

Non-Dispersive

Infrared CO 6000 ppm 257 ppm CO

3.5.5. LICOR 840A

A LICOR 840A, a non-dispersive infrared analyzer, was used to measure the CO₂ and

H₂O concentration at the outlet of the rotameter of the isokinetic sampling setup. In previous

laboratory experiments on gasoline fueled engines, the LICOR 840A was used at the outlet of the

EEPS. This is done to record the CO₂ concentration so that the dilution ratio of the EEPS setup

could be determined. However, as discussed in section 3.5.2, the LICOR 840A could not be used

to determine the dilution ratio because the engine used in these experiments produces negligible

amounts of CO₂.

For these experiments, the LICOR 840A is used to measure the undiluted CO₂

concentration of the exhaust to determine the lubricating oil consumption rate. The LICOR 840A

is incapable of measuring samples above 45°C, so the sample needs to be cooled before being

measured. Since the sample needs to be cooled below 100°C, the water also needs to be removed

before measurement. The CO₂ is measured by the LICOR 840A and then the CO₂ concentration

is adjusted back to a raw exhaust gas basis, with water, using the calculations in Appendix 9.5.

44

The CO₂ concentration could be corrected assuming that all of the water was removed, but a

more accurate method takes into account the water that remains in the sample after the heat

exchanger and desiccant. The LICOR 840A measures the H₂O concentration in addition to the

CO₂ concentration. This H₂O concentration is used in the calculations to correct the LICOR’s

CO₂ measurement to a raw exhaust gas basis.

3.5.6. AFRecorder 2400

The AFRecorder 2400 made by ECM measures the oxygen concentration in the exhaust

and given the fuel type used, which in this case is hydrogen, computes the equivalence ratio at

which the engine is operating.

All of the other emissions equipment is used to collect data to analyze after the test run.

The AFRecorder is different than the other emissions equipment because the equivalence ratio

that it measures is sent to the ECU to control the engine. The equivalence ratio of the engine is

controlled according to an automated closed loop control strategy that uses the equivalence ratio

from the AFRecorder as the feedback signal. The duration of fuel injection, which is metered by

the ECU, is altered to keep the equivalence ratio of the engine constant.

Data Acquisition 3.6.

Labview was used to record most of the data for the experiments. A National Instruments

data acquisition system was used to convert the analog voltages from the engine sensors to

digital signals that the computer could interpret. A National Instruments cDAQ-9178 was used as

a hub for the three analog-to-digital NI modules. Two NI 9211 modules were used for the

thermocouples, and a NI 9205 module was used for the 0-5 volt analog signals.

The calibrations for the pressure, temperature, and torque sensors were determined in-

house using other professionally calibrated high precision sensors. The calibration procedures for

these sensors can be found in Appendix 9.13.

45

The engine has several Manifold Absolute Pressure (MAP) sensors installed at various

locations in the intake system. Each of the MAP sensors records the absolute pressure and

temperature at one of the stages. The Labview program measures and records the following:

MAPs

o At the inlet of the supercharger

o At the Venturi in the inlet of the supercharger

o At the outlet of the supercharger

o At the outlet of the intercooler

Engine torque

Hydrogen fuel flow rate

Emissions Bench

o Total hydrocarbon count

o Oxides of nitrogen

o Carbon dioxide

o Oxygen

o Carbon monoxide

Cylinder head temperature

Oil temperature

Mass air flow sensor temperature

Throttle position sensor

Thermocouples

o Exhaust manifold

o Inlet of catalyst

o Inlet of sampler tube

o Outlet of sampler tube

o Isokinetic probe

o Dynamometer water outlet

46

Throttle Body Controller 3.7.

The original throttle body for this engine was broken when the University of Toronto

received the engine from Ford. Therefore, a replacement throttle body was purchased and fitted

to the engine. A mechanical throttle body and powerful stepper motor were used instead of a

traditional DC powered drive by wire throttle body. Although bulkier, the custom made

mechanical throttle body and stepper motor combination allows for far more precise control than

the traditional drive by wire throttle body.

To control this custom throttle body arrangement, a stepper motor driver controlled by an

Arduino was used. The Arduino code for this stepper motor driver control can be found in

Appendix 9.10.

A Proportional Integral Derivative (PID) closed loop control system for the throttle body

would have been preferred to the fire and forget system used in the end. Several programs were

written to control the throttle body with a PID algorithm using the throttle position sensor as the

feedback signal. Unfortunately, the electrical signal noise on the throttle position sensor channel

made the PID control algorithm ineffective. After several attempts, the PID program was

dropped in favour of a fire and forget system that was more stable and yielded better stability at

the road load condition.

Dynamometer 3.8.

A Go Power DA 312 dynamometer coupled with a Digalog 1022A dynamometer

controller was used to keep the engine at a constant speed as the engine load was increased. A

pneumatically controlled water valve was used to meter the amount of water flowing into the

dynamometer housing. By increasing the amount of water in the dynamometer, blades in the

dynamometer must spin through more water which requires more power. An electronic pressure

transducer on the housing of the dynamometer is used to record the torque produced by the

engine. The calibration procedure for the torque/pressure sensor can be found in Appendix 9.13.

47

Electronic Control Unit 3.9.

A fully customizable Electronic Control Unit (ECU) was purchased from Performance

Electronics. The customizable ECU allowed the experimenters to run the engine at different

equivalence ratios and with various spark timings. The equivalence ratio was sustained at a

constant value with a closed loop control strategy that measured the wideband oxygen

concentration in the exhaust.

The ECU was programed through a software package provided by Performance

Electronics. A desktop computer was used to track parameters of the engine during tests.

Moreover, the ECU stored engine data from the tests which was retrieved after the experiments.

The ECU recorded several important engine parameters including:

Speed

Intake manifold temperature

Intake manifold pressure

Coolant temperature

Mass air flow rate

Fuel open time

Fuel injection angle

Ignition timing

Equivalence ratio

48

4. Methodology and Experimental Procedure

The testing methodology can be separated into three distinct categories. First, spark timing tests

were performed to determine the optimum timing over a range of conditions. Next, a mixture of

supercharged and naturally aspirated tests with fixed spark timing were performed. The overall

test procedure and case specific protocol will be discussed in the upcoming sections.

Test Matrix 4.1.

The order and dates of the tests are shown in Table 4. The spark timing tests were

performed first to assess the optimal spark timing to use for the remaining tests. After the spark

timing tests, the supercharged and naturally aspirated tests were interspersed in an attempt to

mitigate uncontrolled testing factors (i.e. room humidity, etc.). For the same reason, the testing

order of the various equivalence ratios was also randomized.

Table 4. Test Matrix

Date Equivalence

Ratio

Supercharged or Naturally

Aspirated Type of Test

February 11, 2016 0.6 Supercharged Spark Timing Test



February 26, 2016 0.4 Supercharged Fixed Spark Timing



March 3, 2016 0.5 Supercharged Fixed Spark Timing




March 5, 2016 0.6 Naturally Aspirated Fixed Spark Timing





49

General Test Protocol 4.2.

By and large, most of the operating procedure was the same for the different test types

regardless of the specific parameter being tested. In general, the methodology used for the tests is

described below. Any specific differences made to the procedure for individual tests will be

discussed in later sections.

1. The engine battery is charged overnight preceding a test.

2. The next day, the battery charger is removed and the battery voltage is tested.

3. All of the exhaust analyzers discussed in section 3.5 are calibrated. Due to the length of

the calibration procedure, the comprehensive calibration protocol can be found in

appendix 9.12. The procedure for the emissions equipment does not change based on the

test case.

4. The air supply for the pneumatically controlled dynamometer water valve is opened.

5. The computer programs are opened for the EEPS, ECU, Labview, and Arduino

controlled throttle body.

6. The exhaust fan in the room is turned on.

7. The fuel system is purged with nitrogen.

8. All of the emissions equipment is turned on and set to sample room air to get a ten minute

background reading.

9. Valves are actuated for all of the emissions equipment so that they are measuring from

the exhaust.

10. All of the water valves that supply the dynamometer, coolant heat exchanger, intercooler,

and isokinetic probe system heat exchanger are turned on.

11. The fuel rail is filled with hydrogen.

12. The engine is turned on and left to idle for ~1 minute.

13. The load is increased by opening the throttle body (actuated through the Arduino code

shown in Appendix 9.10) until the torque reaches the road load setting. The speed of the

engine is maintained during this process by the dynamometer controller.

14. Once the engine is at the road load condition, the filter is inserted into the isokinetic

probe setup.

50

15. The engine is left to run at the road load condition. During this process, the coolant

temperature is maintained at 80°C with a needle valve at the outlet of the coolant heat

exchanger.

16. After a predetermined length of time (depending on the test type), the filter is removed

from the isokinetic setup and the engine’s load is decreased by closing the throttle body.

17. Once the engine is back at the idle condition, a large quarter turn ball valve is opened at

the outlet of the coolant heat exchanger. This is done to cool down the engine to prepare

for the next test.

18. After idling the engine for five minutes to cool it, the fuel rail is purged with nitrogen to

shutoff the engine.

19. The hydrogen supply is shutoff and purged with nitrogen.

20. All of the emissions sampling equipment is turned off as per the procedure in Appendix

9.12.

Specific Testing Protocol for Spark Timing Tests 4.3.

The purpose of the spark timing tests was to determine the optimal point in the engine’s

cycle to fire the spark plugs. If the spark timing is too advanced, meaning that the spark is fired

far before the piston reaches Top Dead Center (TDC), the NOx emissions will be undesirably

high. If the spark timing is too retarded, meaning that the spark is fired close to TDC, power

output and thus the fuel conversion efficiency will be low. There is no definitive optimal point in

this trade off, but in general, a spark timing that is slightly retarded of the Maximum Brake

Torque (MBT) timing is used as the best compromise between power (fuel conversion

efficiency) and NOx.

For the spark timing tests, the engine was run once for each of the three equivalence

ratios (0.4, 0.5, and 0.6). For all of the spark timing tests, the engine was operated with the

supercharger installed.

The engine was turned on and set to the road load condition. The spark timing was

adjusted and then left for two minutes to stabilize before being adjusted again. The number of

51

spark timing conditions tested varied with the different equivalence ratios. The reason for this is

the difference in safe operating ranges given the different equivalence ratios. If the spark timing

is too advanced, knocking occurs. If the spark timing is too retarded, backfire occurs. As the

equivalence ratio increases, the safe operating range for the spark timing gets smaller. Therefore,

at higher equivalence ratios, fewer spark timing conditions are tested. The information on spark

timing for hydrogen fueled engines in (Natkin, et al., 2003) and (Tang, Kabat, Natkin, &

Stochhausen, 2002) was used to guide the spark timing tests.

For the spark timing tests, no filter samples were taken because the conditions of the

engine were changing with the varying spark timing settings which meant that there was

insufficient time to collect a sample. Additionally, it is worthwhile to point out that the throttle

body position and therefore the intake manifold pressure remained constant throughout the spark

timing tests. As the spark timing was changed, the equivalence ratio and intake manifold

pressure were held constant. This means that the fuel flow rate did not change. However, the fuel

conversion efficiency did change with different spark timings because of the decrease in engine

torque.

Specific Testing Protocol for Naturally Aspirated Tests 4.4.

The naturally aspirated tests are the longest tests performed because the engine has the

highest fuel conversion efficiency without the supercharger. For all of the naturally aspirated

tests, the filter is left in the filter holder in the isokinetic sampling system for 30 minutes. The

engine was operated naturally aspirated with an equivalence ratio of 0.6 for three separate tests.

The engine was not operated naturally aspirated at the other equivalence ratios (0.4 and 0.5).

Additionally, given that the engine was operated without the supercharger for the

naturally aspirated tests, there was no intercooler installed. As a result, for all of the naturally

aspirated tests, the water valve discussed in the general engine operating procedure that feeds the

intercooler was not opened.

52

Specific Testing Protocol for Supercharged Tests 4.5.

The supercharged experiments consist of nine tests. Three tests were performed at each of

the equivalence ratios of 0.4, 0.5, and 0.6. Because the supercharger requires so much power to

operate, the fuel conversion efficiency of the engine is lower. As a result, the supercharged tests

cannot be as long as the naturally aspirated tests due to the fixed quantity of fuel available. For

all of the supercharged tests, the filter was left in the filter holder in the isokinetic system for 20

minutes.

Filter Sample Weights 4.6.

All of the filters used in the experiments were Pall Life Sciences Teflo 47mm 2.0μm

filters. They were weighed with a Sartorius SE 2-F balance in a class 100 clean room. All of the

filters were weighed three times prior to testing and three times after testing. The measurement

procedure for the clean room can be found in Appendix 9.11.

In an attempt to recreate the testing conditions without the engine running, two additional

filter samples were taken. A clean filter was inserted in the filter holder and then the filter holder

was placed inside of a kiln at 250°C to recreate the temperature conditions of the tests. With the

engine off, another filter was inserted in the filter holder and then placed in the isokinetic probe

setup. Without turning the engine on, the isokinetic system pump was turned on to draw a sample

from the exhaust into the filter. Weight tests were not performed on these two additional tests.

Instead, a visual inspection was performed to assess whether any particles were deposited on the

filters.

53

5. Results

All of the tests listed in the text matrix (Table 4) were carried out at a road load condition

simulating operation of a Ford Ranger at 100 km/hr on a level highway. This consisted of a load

(torque) of 82.78 N*m at an engine speed of 2314 RPM for a total brake horsepower of 20.06

kW. For more information about the road load operating condition, see Appendix 9.8.The results

section is separated into six segments:

1. spark timing tests,

2. steady state tests,

3. acceleration period during start-up,

4. lubricating oil consumption rates,

5. gravimetric filters, and

6. emissions equipment.

The tests themselves were performed in two primary stages. The first stage of tests

consisted of running the engine at the road load condition once for each of the equivalence ratios

(0.4, 0.5, and 0.6) and changing the spark timing. These tests were performed first so that the

spark timing could be optimized for the rest of the steady state tests.

The second stage of tests consisted of running the engine at each of the equivalence ratios

(0.4, 0.5, and 0.6) with the supercharger at a fixed spark timing. Each of these conditions was

repeated three times to produce a more statistically significant data set. The engine was also run

three times naturally aspirated at an equivalence ratio of 0.6. The four different test conditions of

this 12 run test matrix (nine supercharged and three naturally aspirated), were interspersed to

create a near random testing sequence. Please refer to Table 4 for the more detailed test order.

For all of the graphs in the results section, the error bars represent two standard

deviations unless otherwise stated.

54

Spark Timing Tests 5.1.

As previously discussed in section 4.3, the spark timing tests were carried out once for

each of the three equivalence ratios. The optimal spark timing for engines is a compromise

between power (fuel conversion efficiency) and exhaust emissions. In general, engines have a

maximum power output for every operating condition (speed and intake manifold pressure) that

is a function of spark timing. The spark timing for the maximum brake torque (MBT) condition

depends on several complicated engine parameters. As a result, even large engine manufacturers

find the MBT timing for new engines through experimental testing.

If the spark timing is advanced or retarded from this maximum power point, the engine’s

power will slowly decrease. Unfortunately, this MBT timing point typically falls in an exhaust

emissions region that is undesirable. Therefore, a balance must be struck between the power

output of the engine and the exhaust emissions for each operating condition. For these

experiments, one operating condition for three equivalence ratios was tested.

For hydrogen fueled engines, the only exhaust emission that is significantly affected by

the spark timing is NOx. As a result, the effect of spark timing on fuel conversion efficiency and

NOx is graphed together for each equivalence ratio to display the trade-off.

55

Figure 16. NOx and Fuel Conversion Efficiency vs. Ignition Timing for φ = 0.4

Figure 16 shows the expected trade-off between fuel conversion efficiency and NOx for

an equivalence ratio of 0.4. The blue line, which corresponds to the left axis, shows the effect of

ignition timing on NOx. The orange line illustrates the relationship between ignition timing and

fuel conversion efficiency; right axis. The bottom axis is the ignition timing and the units are in

degrees before top dead center (°BTDC). All of the green data points represent 2 minute means

from the spark timing tests. The red data points represent 20 minute means from the steady state

supercharged tests which will be discussed in further detail in section 5.2.

As the ignition timing is advanced, the NOx emissions increase. As the spark timing is

advanced, i.e. the spark is fired farther before TDC, the in-cylinder temperatures get higher. This

0

5

10

15

20

25

30

35

0

20

40

60

80

100

120

140

160

180

200

-5 5 15 25 35 45Fu

el C

on

vers

ion

Eff

icie

ncy

(%

)

Emis

sio

ns

Be

nch

NO

x (p

pm

)

Ignition Timing (°BTDC)

Emissions Bench NOx (ppm) and Fuel Conversion Efficiency (%) vs. Ignition Timing (°BTDC) for Spark Timing Tests and Full Tests at φ = 0.4

N0x Spark Timing Tests SC

N0x Full Tests SC

η_fc Spark Timing Tests SC

η_fc Full Tests SC

56

increase in in-cylinder temperatures is a result of the combustion process starting before the

compression stroke is finished. NOx production is heavily dependent on temperature and as a

result, as the spark timing is advanced, the NOx emissions increase.

Figure 16 also shows the effect of spark timing on fuel conversion efficiency with the

orange line. There is a clear maximum in fuel conversion efficiency around 20 °BTDC. The fuel

conversion efficiency and power of the engine are strongly affected by the spark timing. If the

spark timing is too early in the cycle (advanced), the combustion event will work against the

compression process and decrease the power output of the engine. However, if the spark timing

is too late (retarded), the combustion process will happen later in the expansion stroke where less

work can be extracted. A balance needs to be struck between these two extremes which yields

the maximum power output.

For this equivalence ratio of 0.4, the NOx emissions are fairly low throughout the range.

At the MBT timing, the NOx emissions are ~10 ppm. Therefore, for this equivalence ratio, the

spark timing to run the remaining experiments was chosen to be 20 °BTDC.

The red points on Figure 16 show the 20 minute mean from the full supercharged tests.

These points are put on the spark timing graph to illustrate the validity of the spark timing tests.

Even though each condition in the spark timing tests is performed only once for a two minute

period, the agreement of the three longer tests shows that the rest of the graph is valid for

decision making purposes.

57


Figure 17 shows the trade-off between NOx (blue) and fuel conversion efficiency

(orange) for the equivalence ratio of 0.5. For this equivalence ratio, the MBT timing appears to

be ~15 °BTDC. The NOx emissions are significantly higher for the equivalence ratio of 0.5,

which is expected due to the higher combustion temperature. Notice that the left axis for an

equivalence ratio of 0.4 (Figure 16) was zero to 200 ppm whereas the axis for an equivalence

ratio of 0.5 (Figure 17) is zero to 700 ppm.

As a result, for an equivalence ratio of 0.5, a more retarded spark timing of 12.5 °BTDC

was chosen as the optimum compromise between fuel consumption and NOx.

0

5

10

15

20

25

30

35

0

100

200

300

400

500

600

700

-15 -5 5 15 25 35Fu

el C

on

vers

ion

Eff

icie

ncy

(%

)

Emis

sio

ns

Be

nch

NO

x (p

pm

)




N0x Full Tests SC


η_fc Full Tests SC

58


For an equivalence ratio of 0.6, the safe spark timing range for hydrogen is significantly

narrower. As a result, fewer spark timing points were tested for this condition. As the fuel

conversion efficiency line (orange) shows in Figure 18, the MBT timing for this equivalence

ratio is ~12.5 °BTDC. For this equivalence ratio, the NOx emissions are even higher (maximum

y-axis value is 1800 ppm) due to the higher combustion temperatures as the fuel-air equivalence

ratio is increased. The optimal spark timing for this operating condition is less obvious than for

the other equivalence ratios. A spark timing of 7.5 °BTDC was chosen for the full tests to reduce

the NOx emissions without significantly affecting the fuel conversion efficiency.

0

5

10

15

20

25

30

35

0

200

400

600

800

1000

1200

1400

1600

1800

-7.5 -2.5 2.5 7.5 12.5 17.5Fu

el C

on

vers

ion

Eff

icie

ncy

(%

)

Emis

sio

ns

Be

nch

NO

x (p

pm

)




N0x Full Tests SC


η_fc Full Tests SC

59

Figure 19. Available Turbocharger Power vs. Engine Power for Spark Timing Tests

It is important to point out that the ignition timing effects more than just the engine’s

output power. Although the engine for these experiments is operated with a supercharger, a

turbocharger could also be used. The major advantage of turbochargers is that they operate from

the waste energy of the exhaust. As a result, the parasitic losses of turbochargers are far less than

superchargers. This means that the fuel conversion efficiency of a turbocharged engine is higher.

As the spark timing changes, the temperature of the exhaust is greatly affected. If the

spark timing is less advanced, closer to TDC, the exhaust will be hotter. The expansion process

cools the combustion products, so a combustion process that occurs later in the expansion

process is hotter because it has expanded less. Turbochargers need high exhaust temperatures to

0

2

4

6

8

10

12

14

16

18

0 5 10 15 20 25

Ava

ilab

le T

urb

och

arge

r P

ow

er

(kW

)

Engine Power (kW)

Available Turbocharger Power (kW) vs. Engine Power (kW) for Spark Timing Tests

Phi 0.4

Phi 0.5

Phi 0.6

60

drive the air compression process. Therefore, in addition to the compromise between fuel

conversion efficiency and NOx emissions, for turbocharged engines, the exhaust temperature

also needs to be taken into consideration.

The available turbocharger power in the exhaust gas can be estimated from the

experimentally measured temperatures and was calculated assuming a constant turbocharger

outlet temperature. The full calculation for the available turbocharger power can be found in

Appendix 9.7. Figure 19 shows the available turbocharger power vs. engine power for the three

different equivalence ratios (0.4, 0.5, and 0.6). As Figure 17 illustrates, as the ignition timing is

changed to optimize the engine power, the power available to drive the turbocharger is

substantially decreased. It is evident from this result that choosing the optimum spark timing for

a turbocharged engine would be significantly more complex.

Steady State Tests 5.2.

As previously discussed in section 2.2.3, there are several possible operating strategies

for hydrogen fueled internal combustion engines. One operating strategy is to enrich the

equivalence ratio to increase the engine’s power output when demand changes.

Figure 20 illustrates the effect of increasing the equivalence ratio on NOx at a constant

power output; the road load condition described in Appendix 9.8. The power output and speed of

the engine is the same for all of the equivalence ratios. For each of the difference equivalence

ratios, the amount of fuel being burned is relatively constant. However, the amount of air flowing

through the engine is altered by opening the throttle body’s plate more or less which changes the

equivalence ratio.

With the equivalence ratio so far below one, exhaust aftertreatment is difficult. As a

result, there is a huge advantage of running at lean equivalence ratios for part load operation.

However, when the engine reaches the wide open throttle (WOT) condition, the equivalence

ratio needs to be enriched to increase the power output.

61

Figure 20. NOx vs. Equivalence Ratio at the Road Load Setting

Each of the red data points represents a 20 minute mean for one of the supercharged test

at an equivalence ratio of 0.4, 0.5, or 0.6. The three blue data points are 30 minute mean averages

from the three naturally aspirated tests at an equivalence ratio of 0.6. The NOx emissions for the

naturally aspirated tests are slightly lower than the supercharged tests at the same equivalence

ratio, speed, and power output setting. The reason for this is the mechanical power required to

run the supercharger. The peak in-cylinder temperature for the naturally aspirated tests is lower

because there is less air and fuel on a mass basis in the cylinder than the supercharged tests. The

additional air and fuel in the supercharged tests increases the pressure in the cylinder at TDC

which increases the peak in-cylinder temperature.

0

100

200

300

400

500

600

0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 0.7

Emis

sio

ns

Be

nch

NO

x (p

pm

)

Emissions Bench O₂ Calculated Equivalence Ratio

Emissions Bench NOx (ppm) vs. Emissions Bench O₂ Calculated Equivalence Ratio

Supercharged Tests

Naturally Aspirated Tests

62

Figure 21. NOx Produced per km vs. Equivalence Ratio at the Road Load Setting with Emissions

Regulation Comparisons (United States Environmental Protection Agency, 2014) (Johnson,

2014) (MECA, 2014)

It is not possible to estimate vehicle emissions from a single steady-state operating

condition. Vehicle emission certification is based on measurements taken over a driving cycle

that includes many operating conditions. Nonetheless, an order of magnitude comparison can be

made. The emissions at the road load test condition, which were are for a simulated highway

cruise at 100 km/hr, can be used to calculate the mass of emissions on a per km basis. The

specifics of this calculation can be found in Appendix 9.4.

1

10

100

1000

0.3 0.4 0.5 0.6 0.7

NO

x P

rod

uce

d p

er

km (

mg·

NO

x/km

)


NOx produced per km (mg·NOx/km) vs. Emissions Bench O₂ Calculated Equivalence Ratio

Supercharged Tests


EPA Tier 3 (by 2025)

California SULEV20 (by 2025)

EPA Emission Standard (2017)

63

Figure 21 shows the calculated mass of NOx produced per km for each of the tests. It

should be pointed out that the y-axis is a logarithmic scale. Three emissions standards are also

presented in Figure 21. The strictest NOx emissions standard in the United States is the

California Super-Ultra Low Emissions Vehicle 20 (SULEV20) which is shown on Figure 21 in

green. The EPA Tier 3 NOx emission standard is also shown in Figure 21; orange. Both of these

regulations represent the final phase of their respective emissions standards plans which go into

effect in 2025. The purple line shows the EPA regulated emission standard in 2017.

The supercharged (red) and naturally aspirated (blue) data points put this engine’s NOx

emissions in perspective. At an equivalence ratio of 0.4, the engine is capable of beating the

current EPA NOx emissions regulations without exhaust aftertreatment. This is a tremendous

accomplishment. Exhaust aftertreatment devices represent a substantial portion of current

powertrain costs. This means that many of the increased costs associated with moving from a

gasoline fueled engine to a hydrogen fueled engine could be mitigated by saving money on the

exhaust aftertreatment devices.

Another point of comparison is the gasoline direct injection (GDI) engine currently under

test in the same test cell. At a similar road load condition, it produces engine-out NOx emissions

of approximately 2000 ppm (Ramos, 2015). Assuming a catalytic converter efficiency of 90%,

the catalyst-out NOx emissions would be on the order of 200 ppm. The NOx emissions from the

supercharged hydrogen fueled engine at an equivalence ratio of 0.4 are well below this level

without exhaust aftertreatment. The GDI engine equipped with its catalytic converter meets

current EPA vehicle emission standards.

64

Figure 22. Fuel Conversion Efficiency vs. Equivalence Ratio for Supercharged and Naturally

Aspirated Tests at the Road Load Power Setting

Figure 22 shows the effect of equivalence ratio on fuel conversion efficiency for the

supercharged (red) and naturally aspirated tests (blue). Theoretically, running at lower

equivalence ratios should increase the fuel conversion efficiency. The fuel conversion efficiency

is a ratio of the power produced by the engine divided by the power available in the fuel being

burned.

It may seem counterintuitive that the fuel conversion efficiency increases as the

equivalence ratio decreases. It may seem that by running at an equivalence ratio of one, the

maximum amount of power would be produced, which would increase the fuel conversion

0

5

10

15

20

25

30

35

40

45

0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 0.7

Fue

l Co

nve

rsio

n E

ffic

ien

cy (

%)


Fuel Conversion Efficiency (%) vs. Emissions Bench O₂ Calculated Equivalence Ratio

Supercharged Tests


65

efficiency. Although running the engine at an equivalence ratio of one produces the maximum

amount of power for that speed and intake manifold pressure, it also uses a lot fuel.

As the equivalence ratio is decreased, the engine requires more air at the same power and

speed setting. To accomplish this, the throttle plate is opened to increase the pressure of the

intake manifold. By opening the throttle plate, the engine needs to do less work during the intake

process to suck in the incoming charge of air and fuel. By reducing the intake work, the fuel

conversion efficiency of the engine is increased.

Furthermore, by decreasing the equivalence ratio, the in-cylinder temperature is

decreased. The additional air present in the combustion chamber at low equivalence ratios acts

like a diluent that keeps the combustion chamber cooler. Power is extracted from the combustion

process by converting the pressure rise in the cylinder to the linear motion of the piston. If the

burned gases are cooled after the combustion process, the pressure decreases which reduces the

amount of power that can be extracted. When the equivalence ratio is increased, the contents of

the combustion cylinder are hotter which means that there is a larger temperature gradient from

the cylinder contents to the engine block. Therefore, as the equivalence ratio is increased, the

heat transfer out of the combustion chamber increases which decreases the fuel conversion

efficiency.

Finally, by decreasing the equivalence ratio, the specific heat capacities of the burned

gases decrease. By lowering the specific heat capacities of the burned gases, the expansion

process occurs over a larger temperature gradient and thus extracts more power. Although this

effect is not evident at first glance, it has a strong influence on the fuel conversion efficiency and

can be predicted using thermodynamic modeling.

For completeness, there is an effect of decreasing the equivalence ratio that acts to

decrease the fuel conversion efficiency. As the equivalence ratio is decreased, the flame speed

decreases which increases the number of crank angle degrees that the combustion process occurs

over. Although this makes the engine operation smoother, it can also reduce the efficiency. To

optimize the power extracted from the combustion chamber, the pressure rise should happen as

quickly as possible just after TDC. However, hydrogen has an extremely fast flame speed, so this

66

is less of an issue for hydrogen fueled engines. Moreover, the spark timing can normally be

advanced to compensate for this decrease in flame speed.

Although the theory suggests that running at lower equivalence ratios should increase the

fuel conversion efficiency, Figure 22 shows that there is no statistically significant effect of

equivalence ratio on fuel conversion efficiency for the tests. There are likely three reasons for

this:

1. the large parasitic loss of the supercharger,

2. the operating condition for the tests is at part load, and

3. even the richest equivalence ratio tested is fairly lean.

The supercharger requires a significant amount of power to turn which greatly reduces

the fuel conversion efficiency of the engine. Other aspects that affect the fuel conversion

efficiency are dwarfed by the supercharger’s drain on the system.

All of the tests are performed at the road load condition which is heavily throttled even

for the leanest equivalence ratio. However, because of the supercharger, the power required

during the intake process is not a significant concern. Even though the engine is heavily

throttled, the intake manifold is above atmospheric pressure for all of the equivalence ratios. As a

result, the power required by the intake process is fairly low. However, the power required by the

supercharger is substantial because it is essentially pumping against a closed valve (the throttle

plate). Theoretically, at leaner equivalence ratios, the engine would be less throttled and

therefore the supercharger would do less work. Since all of the equivalence ratios are heavily

throttled for the supercharged tests, the significance of running leaner and reducing the pumping

work of the supercharger is negligible.

Finally there is little effect of equivalence ratio on fuel conversion efficiency for the tests

because even the richest equivalence ratio is fairly lean. Possibly the most significant effect of

the equivalence ratio on the fuel conversion efficiency is the impact on the specific heats of the

burned gases. The richest operating point is an equivalence ratio of 0.6 which is quite lean. The

burned gases of the richest operating condition are likely already so diluted that the effect of

further dilution is marginal.

67

The last interesting finding from Figure 22 is the difference between the fuel conversion

efficiency of the supercharged and naturally aspirated tests at an equivalence ratio of 0.6. It is

clear from the graph that a significant portion of power, and thus fuel, is going to power the

supercharger.

The power consumed by the supercharger was estimated using two different

methodologies, as described in Appendix 9.9. The two calculations resulted in supercharger

works of 4.9 kW and 6.2 kW. The second method, which is based on a Willan’s line analysis,

also provides an estimate of the frictional power of the engine; which is 13.1 kW. This is the

power required to overcome the internal friction of the engine and the dynamometer at the test

speed of 2314 RPM. The power produced by the gas acting on the piston is the indicated power.

It is the sum of the brake power and the frictional power.

𝑃𝑖𝑛𝑑𝑖𝑐𝑎𝑡𝑒𝑑 = 𝑃𝑏𝑟𝑎𝑘𝑒 + 𝑃𝑓𝑟𝑖𝑐𝑡𝑖𝑜𝑛

For the naturally aspirated engine, the indicated power is:

𝑃𝑖𝑛𝑑𝑖𝑐𝑎𝑡𝑒𝑑,𝑁𝐴 = 20 𝑘𝑊 + 13.1 𝑘𝑊 = 33.1 𝑘𝑊

For the supercharged engine, the friction power is the sum of the frictional power of the

engine and the power required to drive the supercharger:

𝑃𝑓𝑟𝑖𝑐𝑡𝑖𝑜𝑛,𝑠𝑢𝑝𝑒𝑟𝑐ℎ𝑎𝑟𝑔𝑒𝑑 = 𝑃𝑓𝑟𝑖𝑐𝑡𝑖𝑜𝑛,𝑒𝑛𝑔𝑖𝑛𝑒 + 𝑃𝑠𝑢𝑝𝑒𝑟𝑐ℎ𝑎𝑟𝑔𝑒𝑟 = 13.1 + 6.2 = 19.3 𝑘𝑊

68

The indicated power for the supercharged engine is:

𝑃𝑖𝑛𝑑𝑖𝑐𝑎𝑡𝑒𝑑,𝑠𝑢𝑝𝑒𝑟𝑐ℎ𝑎𝑟𝑔𝑒𝑑 = 20 𝑘𝑊 + 19.3 𝑘𝑊 = 39.3 𝑘𝑊

Therefore, combustion in the supercharged engine has to produce 18.7% more indicated

power than the naturally aspirated case. This is the main reason for the difference in fuel

conversion efficiency between the supercharged and naturally aspirated tests at an equivalence

ratio of 0.6; shown in Figure 22. It also explains the difference between the NOx emissions of

the supercharged and naturally aspirated NOx emissions: shown in Figure 20. Producing more

indicated power requires more fuel, which also produces higher in-cylinder temperatures, a key

factor in NOx formation.

Figure 23. Percentage of NO or NO₂ that Contributes to NOx vs. Equivalence Ratio

0

20

40

60

80

100

120

0.4SC

0.5SC

0.6SC

0.6NA

Pe

rce

nta

ge o

f N

Ox

(%)

Equivalence Ratio and Operational Mode (Supercharged or Naturally Aspirated)

Percentage of NOx Contributed by NO and NO₂ vs. Equivalence Ratio

and Operational Mode (Supercharged or Naturally Aspirated)

NO2

NO

69

Oxides of nitrogen (NOx) are comprised of two species, nitric oxide (NO) and nitrogen

dioxide (NO₂). Figure 23 shows the percentage of NO and NO₂ for each of the equivalence

ratios. As the equivalence ratio decreases, the proportion of NO₂ increases. The reason for this is

the reduction in temperature as the equivalence ratio is decreased (Heywood, 1988). In a

traditional gasoline fueled spark ignition engine, NO is generated in the flame and then quickly

reacted to NO2 through the following type of reaction:

𝑁𝑂 + 𝐻𝑂2 → 𝑁𝑂2 + 𝑂𝐻

In a traditional gasoline fueled spark ignition engine, because of the high temperatures

owed to nearly stoichiometric equivalence ratios, this NO2 is then reacted back to NO through

the following reaction:

𝑁𝑂2 +𝑂 → 𝑁𝑂 + 𝑂2

In a diesel or hydrogen fueled engine operating at low equivalence ratios, the first

reaction from NO to NO2 still occurs. However, due to lower temperatures in the flame region,

the rate of the second reaction slows, which increases the proportion of NO2 in the exhaust.

70

Figure 24. Intake Manifold Pressure vs. Engine Power for Supercharged and Naturally Aspirated

Tests with Different Equivalence Ratios

Figure 24 shows the intake manifold pressure vs. engine power for different equivalence

ratios with the engine supercharged and naturally aspirated. Similarly to the Willan’s line graph

(Figure 37), the data for Figure 24 is from the loading period of the engine at constant speed.

Figure 24 can be used to determine the point where supercharging or increasing the equivalence

ratio is necessary to meet the required power output. As Figure 24 shows, with each of the

equivalence ratios, the power increases fairly linearly as the intake manifold pressure increases.

There is a plateau region for each of the equivalence ratios after the increase in intake manifold

pressure. It appears that the power output of the engine increases without any apparent increase

in intake manifold pressure. This is caused by the reduction in friction as the engine heats up and

0

20

40

60

80

100

120

140

0 5 10 15 20 25 30

Inta

ke M

anif

old

Pre

ssu

re (

kPa)

Engine Power (kW)

Intake Manifold Pressure (kPa) vs. Engine Power (kW) for Supercharged and Naturally Aspirated

Tests

SC 0.4

SC 0.5

SC 0.6

NA 0.6

71

reaches steady state. Moreover, there are several outliers in the data sets that show very low

engine power output with relatively high intake manifold pressure. These data sets are generated

during the very fast loading process of the engine. The outliers are a result of the torque sensor

used to calculate the engine power receiving erroneous readings because of electrical noise

during the loading period.

Acceleration During Start-up 5.3.

Throughout the entire test matrix, no particulate matter emissions were detected by the

EEPS during the steady state road load condition. The same emissions equipment has been used

in other experiments and has been validated with other particle measuring instruments.

Moreover, the EEPS used for these experiments was also used for tests on other engines in

between the test matrix for these experiments. The EEPS correctly identified PM in the exhaust

of the other engine. Therefore, it can be said with a high degree of confidence that the EEPS was

in perfect working condition and that the results it provided are correct for these tests.

Although the EEPS detected no measureable particulate matter emissions during the

steady state road load condition, some of the tests showed statistically significant particulate

matter spikes during the acceleration period of the engine at start-up.

72

Figure 25. 1-Minute PM Average Concentration and Engine Speed vs. Time for Supercharged

Spark Timing Test φ = 0.4 February 11, 2016

Figure 25 shows the 1-minute average of the particulate matter concentration (blue) and

engine speed (red) vs. time. The engine speed is displayed on this graph to show when the engine

was turned on. The 1-minute average of the particulate matter concentration from before engine

start-up is sampling HEPA filtered air. Therefore, the PM concentration from before the engine’s

start-up represents the minimum detection limit of the EEPS. As Figure 25 illustrates, during the

steady state road load condition, no particulate matter is detected. However, there is a large spike

in PM when the engine is turned on. However, as the next graph shows, this spike in particulate

matter is not present in every test. For Figure 25 and Figure 26, the 1-minute average particulate

matter concentrations are on the basis of the diluted mixture. These PM concentrations were not

0

500

1000

1500

2000

2500

3000

0.E+00

1.E+05

2.E+05

3.E+05

4.E+05

5.E+05

6.E+05

7.E+05

8.E+05

9.E+05

1.E+06

0 500 1000 1500 2000 2500 3000

Engi

ne

Sp

ee

d (

RP

M)

1-M

inu

te P

M A

vera

ge C

on

cen

trat

ion

(#/

cm³)

Time (s)

1-Minute PM Average Concentration (#/cm³) and Engine Speed (RPM) vs. Time (s) for

Supercharged Spark Timing Test φ = 0.4 February 11, 2016

1-Minute Average (#/cm³)

RPM

73

converted to the raw exhaust gas basis because the flat PM line in Figure 26 is easier to

understand on the diluted basis. If the PM concentration was converted to the raw exhaust gas

basis for Figure 26, the PM concentration would jump back and forth between a very small

positive and negative number.

Figure 26. 1-Minute PM Average Concentration and Engine Speed vs. Time for Supercharged φ

= 0.4 March 10, 2016

Figure 26 shows PM and speed vs. time for the same equivalence ratio as Figure 25.

Theoretically, Figure 25 and Figure 26 should show virtually the same trend. However, even

though they are repeats of the same operating condition, they do not show the same trend. In

0

500

1000

1500

2000

2500

3000

0.E+00

1.E+05

2.E+05

3.E+05

4.E+05

5.E+05

6.E+05

7.E+05

8.E+05

9.E+05

1.E+06

0 500 1000 1500 2000 2500 3000 3500

Engi

ne

Sp

ee

d (

RP

M)

1-M

inu

te P

M A

vera

ge C

on

cen

trat

ion

(#/

cm³)

Time (s)

1-Minute PM Average Concentration (#/cm³) and Engine Speed (RPM) vs. Time (s) for

Supercharged φ = 0.4 March 10, 2016

1-Minute Average (#/cm³)

RPM

74

Figure 26, there is no PM spike during engine acceleration. Several graphs were produced in an

attempt to explain this sporadic PM spike phenomenon.

Figure 27. Peak 1-Minute PM Average Concentration vs. Peak Engine Acceleration

For the all of the peak 1-minute PM graphs in this section, the PM concentration is given

on a raw exhaust gas basis. The formula used to convert from the diluted basis to the raw exhaust

gas basis is presented in Appendix 9.16.

One theory for the particulate matter spike occurring in some of tests and not in others is

the start-up speed being different. Each time the engine is turned on, the acceleration with which

0.E+00

1.E+07

2.E+07

3.E+07

4.E+07

5.E+07

6.E+07

0 200 400 600 800 1000 1200 1400

Pe

ak 1

-Min

ute

PM

Ave

rage

Co

nce

ntr

atio

n (

#/cm

³)

Peak Engine Acceleration (RPM/s)

Peak 1-Minute PM Average Concentration (#/cm³) vs. Peak Engine Acceleration (RPM/s) on

a Raw Exhaust Gas Basis

SC

NA

75

it turns on is slightly different. It was theorized that the tests which had particulate matter spikes

might have a faster starting acceleration. Faster engine acceleration would affect the sealing of

the piston rings which keep lubricating oil out of the combustion chamber. Figure 27 shows the

peak 1-minute average PM concentration vs. peak engine acceleration. Unfortunately, there is no

correlation between the PM spike and the acceleration with which the engine is turned on.

Figure 28. Peak 1-Minute Average PM Average Concentration vs. Coolant Temperature

Since all of the particulate matter generated from a hydrogen fueled engine is from the

lubricating oil, the temperature of the lubricating oil is highly related to the particulate matter

generated. The coolant temperature can be used as a proxy for the oil temperature during start-up

0.E+00

1.E+07

2.E+07

3.E+07

4.E+07

5.E+07

6.E+07

0 10 20 30 40 50 60

Pe

ak 1

-Min

ute

PM

Ave

rage

Co

nce

ntr

atio

n (

#/cm

³)

Coolant Temperature (°C)

Peak 1-Minute PM Average Concentration (#/cm³) vs. Coolant Temperature (°C) on a Raw

Exhaust Gas Basis

SC

NA

76

because both are in direct contact with the engine block. Figure 28 shows the peak 1-minute PM

average concentration during start-up vs. coolant temperature during start-up. Unfortunately,

Figure 28 shows no correlation between the spikes in PM and coolant temperature.

Figure 29. Peak 1-Minute PM Average Concentration vs. Testing Order of that Day

Another theory for the sporadic PM spikes is the testing order of that day. The theory was

that after the first test, the lubricating oil would be hotter and therefore more likely to evaporate

from the cylinder wall and turn into PM. As Figure 29 shows, there is no relation between the

PM spikes and the testing order of that day.

0.E+00

1.E+07

2.E+07

3.E+07

4.E+07

5.E+07

6.E+07

0 1 2 3 4

Pe

ak 1

-Min

ute

PM

Ave

rage

Co

nce

ntr

atio

n (

#/cm

³)

Testing Order of that Day

Peak 1-Minute PM Average Concentration (#/cm³) vs. Testing Order of that Day on a Raw

Exhaust Gas Basis

SC

NA

77

Figure 30. Peak 1-Minute PM Average Concentration vs. Nominal Equivalence Ratio

The final theory was that the target equivalence ratio was affecting the PM spike. The

fuel and ignition table, the tables used by the engine to dictate fuel injection and ignition timing,

are different for each of the three equivalence ratios. As a result, the engine response during

starting varies by equivalence ratio. It was thought that this might explain the PM spikes in some

of the tests and not others. However, as Figure 30 shows, there is no correlation between

equivalence ratio and PM spikes. Each of the equivalence ratios have at least one test where

there is a PM spike and at least one test where there is no PM spike.

0.E+00

1.E+07

2.E+07

3.E+07

4.E+07

5.E+07

6.E+07

0.3 0.4 0.5 0.6 0.7

Pe

ak 1

-Min

ute

PM

Ave

rage

Co

nce

ntr

atio

n (

#/cm

³)

Nominal Equivalence Ratio

Peak 1-Minute PM Average Concentration (#/cm³) vs. Nominal Equivalence Ratio on a Raw

Exhaust Gas Basis

SC

NA

78

Lubricating Oil Consumption Rate 5.4.

An interesting result from hydrogen fueled engines is the lubricating oil consumption

rate. In hydrocarbon fueled engines, the lubricating oil consumption rate is difficult to measure

accurately. There are three typical methods, each with their own complications:

1. lubricating oil mass difference,

2. radioisotope tracer, and

3. SO₂ tracer.

By running an engine for extended periods of time, the lubricating oil consumption rate

can be calculated by weighing the lubricating oil at the beginning of the test and at the end.

There are several issues with this method. The final mass of the oil is underrepresented because

some of the oil inevitably sticks to the walls. By using the mass difference method, the engine

needs to be run for a very long time to achieve adequate accuracy. The length of the test also

limits the number of conditions that can be realistically tested. Finally, this method is only

capable of assessing the lubricating oil mass flow rate of a steady state condition. The procedure

of this method also makes it impossible to test transient events.

Another testing technique uses radioisotope doped lubricating oil. By measuring the

concentration of the radioisotope tracer in the exhaust, the lubricating oil mass flow rate can be

determined. The main disadvantage of this technique is the difficulty in using radioisotopes.

The most common technique to measure the lubricating oil consumption rate is the SO₂

tracer technique. Engine lubricants contain sulfur based compounds which are converted

primarily to SO₂ when burned in the combustion chamber. As long as the fuel being burned in

the engine has low sulfur levels, measuring the SO₂ concentration in the exhaust can be used to

calculate the lubricating oil mass flow rate. The major disadvantage to this technique is the

accuracy of the measurement when the oil consumption rates are low.

Another much less commonly used option is to run the engine on hydrogen. Hydrogen

fueled engines produce no fuel derived CO₂. Moreover, lubricating oil is primarily composed of

hydrocarbons, so most of the lubricating oil consumption results in CO₂. By measuring the CO₂

concentration in and out of the engine, the lubricating oil consumption rate can be determined.

79

This method is much more accurate that the SO₂ technique, which lends itself to engines with

low lubricating oil consumption rates.

Figure 31. Lubricating Oil Consumption Rate

Figure 31 shows the lubricating oil consumption rate vs. equivalence ratio for the steady

state period at the road load condition. As the graph shows, the lubricating oil consumption rates

are very low. Moreover, there is no apparent relation between the lubricating oil consumption

rate and the equivalence ratio. All of this goes to support the EEPS results in section 5.3 which

showed that there was no statistically significant PM generated. The calculations used to

generate Figure 31 are shown in Appendix 9.3.

-1

0

1

2

3

4

5

0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 0.7Lub

rica

tin

g O

il C

on

sum

pti

on

Rat

e (

g/h

r)


FTIR Calculated Lubricating Oil Consumption Rate (g/hr) vs. Emissions Bench O₂ Calculated

Equivalence Ratio

Supercharged Tests


80

Figure 32. Lubricating Oil Consumption Rates of Various Engine Types (Kapetanovic, Wallace,

& Evans, 2009) (Froelund, Menezes, Johnson, & Rein, 2001)

Figure 32 illustrates the lubricating oil consumption rates of three engines: the hydrogen

fueled engine from these experiments, a 3.8 litre V6 gasoline fueled engine, and a 3.9 litre B3.9-

C Cummins diesel engine. The lubricating oil consumption rates of the three engines have

roughly the same speed and percent load conditions. The graph shows that the engine in these

experiments has an extremely low lubricating oil consumption rate in comparison to other types

of engines. This further validates the low PM emissions from this engine.

0

5

10

15

20

25

30

35

40

Hydrogen Engine Gasoline Engine Diesel Engine

Lub

rica

tin

g O

il C

On

sum

pti

on

Rat

e (

g/h

r)

Lubricating Oil Consumption Rate (g/hr) for Three Different Types of Engines

81

Filter Analysis 5.5.

The filters that were collected during the naturally aspirated and supercharged fixed spark

timing tests will be discussed in this section. All of the filters were collected in the isokinetic

probe setup explained in section 3.5.1.

Figure 33. Mass Collected On Filters

Figure 33 shows the mass collected on the filter elements vs. target equivalence ratio for

the naturally aspirated (blue) and supercharged (red) tests. The collected mass was calculated by

subtracting the mean of the three pretested filter masses from the mean of the three post-tested

0.0000

0.1000

0.2000

0.3000

0.4000

0.5000

0.6000

0.3 0.4 0.5 0.6 0.7

Mas

s C

olle

cte

d (

mg)

Nominal Equivalence Ratio

Mass Collected on Gravimetric Filters (mg) vs. Nominal Equivalence Ratio

SC

NA

82

filter masses. The error bars on the y-axis were generated by adding the two standard deviations

of the three pretested filter masses with the two standard deviations of the three post-tested filter

masses. Therefore, the error bars shown in Figure 33 represent a very liberal accounting of the

range of the ‘true’ mass collected. Even with this very strict assessment of the error bars in

Figure 33, it is clear that a statistically significant mass was deposited on each of the filters in the

tests. This seems to indicate that PM is being generated by the engine during the steady state

road load condition; a conflicting result to the EEPS.

The filters collect something during the tests, that much is clear. However, based on

visual inspection of the filters, the mass collected is not PM. Figure 34 shows a clean unused

filter on the left and a tested filter on the right. Figure 34 is representative of the other filter

elements. Pictures of the filter elements from the rest of the tests can be found in Appendix 9.14.

Figure 34. Clean Filter (Left) and Tested Filter (Right) Naturally Aspirated at an Equivalence

Ratio of 0.4

83

The tested filter (right) does show contaminants on its surface. However, the

contaminants are large flakes localized on the outer perimeter of the filter element. Engines

generate particulate matter that is orders of magnitude smaller than the debris seen on the tested

filter in Figure 34. Based on similar tests performed on other engine setups that do produce PM,

it is known what the filters should look like with engine generated PM deposited on the filter. In

an engine that generates PM emissions, the center of the filter element is uniformly discoloured

gray to black depending on the amount of PM deposited. The PM is far too small to see

distinguishable pieces; it appears to be evenly coated.

Based on this information, the question becomes, what is on the filters? Two additional

tests were conducted in an attempt to answer this question. First, a filter element was placed in

the filter holder and then the filter holder was placed in a kiln at 250°C for 30 minutes. The filter

holder has several internal gaskets, it was thought that at high temperatures these gaskets might

be breaking apart and soiling the filter element. Based on a visual inspection, this test resulted in

no debris being deposited on the filter element. Second, a clean filter element was placed in the

filter holder and then placed in the isokinetic apparatus. The pump in the isokinetic setup was

turned on and a sample was drawn through the filter element with the engine off. Again, based

on a visual inspection, this test resulted in no debris being deposited on the filter element.

Pictures of the filter elements from these two additional tests and all of the experiments in the

test matrix can be found in Appendix 9.14.

Therefore, the debris on the filter elements is a result of the engine being on and gases

flowing through the exhaust system. The tubes that route the exhaust from the exhaust manifold

to the exhaust sampler tube are comprised of several stainless steel tubes bolted together with

flanges. Each of the flanges is mated together with a high temperature stainless steel reinforced

graphite gasket. During installation of the exhaust system, it was noticed that these gaskets have

a tendency to break apart into small graphite flakes. It is very likely that the debris found on the

filters is from these gaskets. Future work should place a cyclone in the isokinetic probe setup to

remove large debris from the sample stream before entering the filter holder.

84

Emissions Equipment 5.6.

The FTIR and emissions bench both measure NOx, which is the only regulated exhaust

emission that hydrogen fueled engines make in any significant quantity. Therefore, the NOx

results from the FTIR and emissions bench can be compared.

Figure 35. FTIR NOx vs. Emissions Bench NOx

Figure 35 shows the FTIR NOx vs. emissions bench NOx for the naturally aspirated

(blue) and supercharged (red) tests. Perfect agreement between the FTIR and emissions bench

would result in a relation of y = x. As Figure 35 illustrates, the relation is y = 1.0006x+1.3849

y = 1.0006x + 1.3849

y = 1.1167x - 23.65

0

100

200

300

400

500

600

0 100 200 300 400 500

FTIR

NO

x (p

pm

)

Emissions Bench NOx (ppm)

FTIR NOx (ppm) vs. Emissions Bench NOx (ppm)

Supercharged Tests


85

and y = 1.1167x-23.65 for the supercharged and naturally aspirated tests respectively. Both of

these relationships between the FTIR and emissions bench NOx show excellent agreement. The

naturally aspirated relationship is worse than the supercharged relationship because of the

smaller sample size and narrower measured NOx range.

Another comparison that can be made is the O₂ measurement from the emissions bench

and the H₂O measurement from the FTIR. With the O₂ concentration, the H₂O concentration can

be predicted with stoichiometry. Similarly, the O₂ concentration can be predicted with the H₂O

concentration. This analysis can be used to compare the O₂ results from the emissions bench with

the H₂O results from the FTIR. Figure 36 shows the molar concentration of water vs.

equivalence ratio.

Figure 36. Water Concentration Measured vs. Theoretical

0

5

10

15

20

25

30

35

40

0.1 0.3 0.5 0.7 0.9

Mo

lar

Wat

er

Co

nce

ntr

atio

n (

%)

Equivalence Ratio

Molar Water Concentration (%) vs. Equivalence Ratio

Hydrogen Theoretical

Methane Theoretical

SC Hydrogen Tests

NA Hydrogen Tests

NA Methane Tests

86

The green line in Figure 36 shows the theoretical relationship between the equivalence

ratio and water concentration as predicted by stoichiometry for hydrogen combustion. The

supercharged data points (red) and naturally aspirated data points (blue) show the H₂O

concentration from the FTIR vs. the equivalence ratio calculated from the emissions bench’s O₂

concentration. As Figure 36 illustrates, the FTIR’s H₂O concentration is consistently below the

level predicted by the emissions bench O₂ calculated equivalence ratio.

The orange line in Figure 36 shows the theoretical relationship between the water

concentration and the equivalence ratio for methane combustion. The purple data points are for

naturally aspirated tests with the engine running on natural gas (Pop, 2016).

The hydrogen and natural gas tests both show under predicted water concentrations by

the FTIR. Water takes up roughly 20% of the exhaust by volume for the hydrogen fueled tests at

an equivalence ratio of 0.6. Even though the exhaust is heated to 191°C before entering the

FTIR, it is likely that some of the water condensed in the sample line and was lost. This would

account for the consistent under representation of water by the FTIR. It is also possible that the

emissions bench O₂ measurement, which was measured ‘dry’, still had a small amount of water

in the sample. If there was still water in the sample, the oxygen concentration would be under

represented. This would cause the calculated equivalence ratio to be higher than the true value.

As the equivalence ratio increases, the expected water concentration increases as Figure 36

shows. It is likely that both of these effects played into the FTIR water concentration being lower

than expected based on the emissions bench O₂ concentration.

87

6. Discussion and Conclusion

The initial intention of the project was to study the effect of lubricating oil composition on PM

morphology. However, over the course of the project, it was found that this modified research

engine produced effectively no PM at the steady state road load condition. This is in itself a

significant finding, in that the results show that it is possible to produce an IC engine that emits

virtually unmeasurable levels of PM emissions without exhaust aftertreatment. The key is to

control the flow of lubricating oil. The flow of lubricating oil should be reduced as much as

possible while minimizing the friction and wear of the engine.

For hydrogen fueled engines, the only environmentally harmful gaseous emission

produced in significant quantities is NOx. If the engine could be operated with low NOx as well

as low PM emissions, it could be an interesting ultra-low emission replacement option for

gasoline fueled internal combustion engines, a lower cost alternative to fuel cells. The current

cost of fuel cells is approximately $55/kW while a gasoline fueled engine is $39/kW (Ogden,

Steinbugler, & Kreutz, 1998) (Office of Energy Efficiency & Renewable Energy, 2016).

Hydrogen fueled engines would only slightly exceed this $39/kW because of the change to the

fuel storage and delivery system; all of the other components remain the same. Moreover, for the

hydrogen fueled engine, no catalytic converter is needed; a considerable savings.

In general, three types of tests were performed. The engine was run at three equivalence

ratios with variable spark timing to generate a spark timing study. The engine was then run with

fixed spark timing supercharged and naturally aspirated for extended periods of time.

The spark timing tests showed the expected trade-off between NOx emissions and

power/fuel conversion efficiency. As the spark timing was advanced, the NOx emissions

increased and the fuel conversion efficiency increased to a maximum before decreasing again.

On the other hand, retarding the spark timing decreased the NOx emissions at the expense of fuel

conversion efficiency. It was found that the MBT timing was 20, 15, and 12.5 °BTDC for the

equivalence ratios 0.4, 0.5, and 0.6 respectively. The NOx emissions were fairly low throughout

the spark timing sweep for the equivalence ratio of 0.4. As a result, the trade-off between fuel

conversion efficiency and NOx emissions for an equivalence ratio of 0.4 was less pronounced.

However, for the equivalence ratios 0.5 and 0.6, the NOx emissions increased substantially

88

towards to the fuel conversion efficiency maximum. Therefore, the ‘optimal’ spark timing point

for these test conditions was less obvious.

For the longer tests, a fixed spark timing of 20, 12.5, and 7.5 °BTDC was chosen for the

equivalence ratios 0.4, 0.5, and 0.6 respectively. These fixed spark timings were chosen by

moving slightly to the retarded side of the fuel conversion efficiency maximum. The aim was to

strike a balance between moderately reducing the fuel conversion efficiency while substantially

reducing the NOx emissions. The spark timing was chosen such that the fuel conversion

efficiency for the full tests is less than 5% lower than the MBT fuel conversion efficiency.

The longer fixed spark timing tests highlighted the differences between the various

equivalence ratio conditions. The average NOx emissions for the supercharged tests were 11.5,

65.9, and 404.3 ppm for equivalence ratios 0.4, 0.5, and 0.6 respectively. This can be converted

to a mass of NOx per km. The NOx emissions are an average of 31.4, 149.5, and 787.0

mg*NOx/km for the equivalence ratios 0.4, 0.5, and 0.6 respectively. The current EPA

regulation is 99.4 mg*NOx/km which will be gradually reduced to 18.6 mg*NOx/km in 2025.

The EPA requirements apply to vehicle emissions over a test cycle. However, these tests at a

single steady-state operating condition suggest that a hydrogen fueled engine can be operated

without exhaust aftertreatment and still meet the current EPA emissions standards. With further

tuning, or leaner operation, it may be possible to operate the engine below the 2025 regulations.

Changing the equivalence ratio did not exhibit any effect on the fuel conversion

efficiency. Theoretically the fuel conversion efficiency should increase as the equivalence ratio

decreases. However, that trend was not observed in these tests. It is hypothesized that the fuel

conversion efficiency did not change with equivalence ratio for three reasons. First, the

supercharger requires so much power to operate that other effects on the fuel conversion

efficiency are overshadowed. Moreover, the richest equivalence ratio for the tests, 0.6, is already

quite lean. This means that the burned gases are already fairly cold and their specific heat is low.

It is possible that further dilution of the burned gases has a negligible effect on the temperature

of the burned gas. Finally, all of the tests were heavily throttled, even for the equivalence ratio of

0.4. The reduction of pumping losses in the intake system is one of the primary advantages of

running the engine lean. By heavily throttling the engine, much of this benefit is erased.

89

The engine was operated both naturally aspirated and supercharged at an equivalence

ratio of 0.6. This allows for several interesting comparisons of the two engine configurations.

The most immediate impact of adding the supercharger is the reduction in fuel conversion

efficiency. The fuel conversion efficiency of the engine is 28.1% supercharged and 35.1%

naturally aspirated. The supercharger used in the tests is a twin-screw supercharger which is a

style of positive displacement pump. The advantage of this type of supercharger is the power

gains at low engine speeds. However, as the data shows, the disadvantage of this type of

supercharger is the significant amount of power it requires to operate. A turbocharger would be

more advantageous on a fuel conversion efficiency basis. However, a much more complicated

turbocharger study would be required to tune the system properly.

The EEPS results show that during the steady state road load condition, no detectable PM

is generated by the engine. Gravimetric filter samples were also taken at the steady state road

load condition. Although a statistically significant mass was collected on all of the filters, as

section 5.5 showed, this mass did not come from PM. In fact, the filters showed no sign of PM

deposits. Both of these results were further substantiated with visual inspection of the inside of

the exhaust system. For hydrocarbon fueled engines, the inside of the exhaust is lined with soot.

For the engine in these experiments, after all of the tests, the inside of the engine was checked for

soot deposits and it was spotless.

Although the engine produces PM emissions far below the normal detection limits during

steady state operation, it does sometimes transiently produce PM emissions during the start-up of

the engine. These PM spikes during engine start-up were only present for some of the tests.

Several factors were studied in an attempt to answer the sporadic nature of the PM spikes. The

PM concentration was graphed vs. engine acceleration, coolant temperature, testing order of that

day, and target equivalence ratio to assess causal relationships. Unfortunately, no correlation was

found between the sporadic PM spikes and any of the variables studied. Currently there is no

explanation for some of the tests exhibiting PM spikes during engine start-up and others with

virtually the same conditions not exhibiting PM spikes.

The steady state lubricating oil consumption rate was calculated for all of the fixed spark

timing tests. The CO₂ and CO measurements from the FTIR were used to calculate the

lubricating oil consumption rates. The LICOR’s CO₂ measurement was used to verify the results

90

from the FTIR. The equivalence ratio and operational mode (supercharged or naturally aspirated)

had no effect on the oil consumption rate. The oil consumption rates for all of the tests were

extremely low. This further substantiates the findings of the EEPS and gravimetric filters. The

engine produces extremely low PM levels at the steady state road load condition.

Additionally, the FTIR and emissions bench results were compared. As section 5.6

showed, the NOx measurements of the FTIR and emissions bench were in excellent agreement

for all of the tests. The FTIR’s H₂O measurements were also compared to the O₂ results from the

emissions bench. The H₂O results from the FTIR were consistently below the level predicted

from the emissions bench O₂ measurements. The reason for this is likely two fold. First, the

emissions bench measures the O₂ on a dry basis, meaning that the water from the exhaust sample

is removed. Although most of the water is removed, it is probable that some water remains in the

sample. This remaining water would cause the O₂ measurement to be lower than the true value.

A lower O₂ measurement results in a higher expected H₂O level. In other words, if the O₂

concentration is under reported, the expected H₂O concentration will be higher. The other

possible reason for the FTIR under reporting the H₂O concentration is condensation. Although

the FTIR sample lines are kept at 191°C, it is probable that some water condenses out of the

sample in the tubing fittings that connect the heated sample line.

In conclusion, a hydrogen fueled spark ignited internal combustion engine was operated

at various equivalence ratios both naturally aspirated and supercharged. Consistent with other

literature in the area, the hydrogen fueled engine produced NOx emissions that varied

significantly with equivalence ratio. At the leanest operating condition, the engine met current

EPA regulations without exhaust aftertreatment. The twin-screw supercharger drastically

reduced the fuel conversion efficiency of the engine while producing high boost levels at low

speeds. The engine exhibited sporadic PM spikes during the acceleration period of the engine at

start-up. Other researchers have observed large PM spikes during acceleration periods, but none

have reported the sporadic nature seen in these tests. The engine also produced no detectable PM

emissions during the steady state road load condition. This result differs significantly from other

research in the field. This research shows that a hydrogen fueled internal combustion engine is

capable of generating virtually no PM emissions with good enough oil control. Moreover,

consistent with other research in the field, if the engine is run lean enough with boost, the

91

hydrogen fueled engine produces nominal levels of regulated gaseous emissions. Therefore, this

research presents an engine which produces extremely low PM, CO, CO₂, and NOx emissions

without exhaust aftertreatment at the steady state road load condition.

The results from these experiments also have far reaching consequences for gasoline

fueled engines. Previous work in the field has shown that lubricating oil consumption contributes

significantly to PM emissions (Miller, Stipe, Habjan, & Ahlstrand, 2007). The general opinion

has been that the composition of lubricating oils will have to be regulated to minimize the PM

contribution. This thesis shows that there is another option. With the right engine design, the

contribution of lubricating oil consumption to PM emissions can be reduced to essentially zero.

92

7. Future Work

There are two distinct paths that this project can take moving forward. Hydrogen fueled engines

have two primary uses. First, hydrogen fueled engines could be used in the future to replace

gasoline fueled engines. Therefore, research is needed to resolve many of the issues surrounding

the operation of a vehicle on hydrogen. Second, hydrogen fueled engines can be used to better

understand the PM formation of gasoline fueled engines. Hydrogen fueled engines produce PM

solely from oil consumption. Gasoline fueled engines produce PM from the fuel as well as the

lubricating oil. As a result, hydrogen fueled engines can be used to isolate the PM formation

from the lubricating oil of traditional gasoline fueled engines.

If the project is directed towards researching the viability of hydrogen fueled engines in

vehicles, a turbocharger study could be the next step. The experiments performed so far have

shown that an engine can be operated on hydrogen to produce extremely low PM, CO₂, CO, and

NOx emissions. However, the fuel conversion efficiency of the supercharged engine used in

these tests was extremely low. For this type of powertrain to be considered as a replacement for

gasoline fueled engines, it needs to convert the hydrogen into power in a much more efficient

manner. The fuel conversion efficiency of the engine in the naturally aspirated mode shows that

this type of engine is quite efficient without the supercharger. A turbocharger should be fitted to

the engine and tests should be conducted to further validate the possibility of this type of engine

setup in a vehicle.

The engine could also be used to study the potential of Exhaust Gas Recirculation (EGR).

By running the engine with EGR, an equivalence ratio of one can be achieved. The theory of this

operational mode is that a TWC could be used to reduce the NOx in the exhaust. However, given

the results from this study, it would seem to suggest that there are no components to oxidize,

making the TWC significantly less efficient. It would be possible to run the engine rich (i.e. with

an equivalence ratio above one) which would generate unburned hydrogen in the exhaust. This

unburned hydrogen would react well in the TWC and allow for the reduction of NOx. However,

onboard storage of hydrogen in vehicles provides a serious space and weight concern. A strategy

that depends on rich operation is likely to be uneconomical. Moreover, with such large quantities

of water in the exhaust, EGR would be difficult. Most of the water would need to be removed

93

before being put into the intake manifold. This would facilitate an extremely expensive and

complicated EGR system.

The other potential direction of the project is to study the effect of lubricating oil

composition on PM morphology. To perform these tests, many alterations should be made to the

engine setup. The current engine setup did not produce PM because of the alterations that were

made by Ford. These alterations by Ford will essentially need to be undone to conduct PM

morphology research. The valve seals, pistons, and connecting rods will need to be restored to

stock factory parts. This will make the PM results from the hydrogen fueled engine setup

contiguous with average vehicles on the road. The results from these tests can be used to assess

the impact of lubricating oil composition on PM morphology.

Another PM emissions study that can be performed with this setup is a test of the oil

coalescing filter in the PCV system. The original test matrix for this thesis included naturally

aspirated tests without the oil coalescing filter. One test was performed in this condition and no

PM was generated. However, it is possible that an excessively long PCV hose resulted in no PM

emissions. In a normal engine setup, the PCV hose would have been much shorter. With a

modest change of the PCV hoses, the influence of the oil coalescing filter on PM emissions

could be studied. Most gasoline fueled engines do not have oil coalescing filters installed. If PM

emissions were observed from these tests, the results could be used to assess the contribution of

PM from the PCV system in a traditional gasoline fueled car.

In addition to these tests, there are several small alterations that should be made to the

test setup before moving forward. The isokinetic probe setup should be altered to include a

cyclone upstream of the filter element. Over the course of these tests, non-PM derived debris was

deposited on the surfaces of the filter elements. The addition of a cyclone will remove this large

debris and produce more meaningful filter results.

There were several issues experienced with the throttle body controller over the course of

testing. A fire and forget algorithm was used for the tests because of simplicity. For future tests,

the throttle body controller should be programed in a PID algorithm. This will allow for more

complex driving cycles to be tested.

94

For more complex driving cycles to be used effectively, the dynamometer controller will

need to be replaced with a faster reacting model. The dynamometer controller used for these tests

was slow to respond to changing loads despite being tuned several times. To mimic the EPA

driving cycles, a dynamometer with a faster response time would be required.

The dilution system used for the EEPS in these tests consisted of a positive displacement

pump and two diluters. The EEPS requires a sample flow rate of 10 L/min. The positive

displacement pump that the company recommends for this setup is incapable of supplying 10

L/min. As a result, the exhaust sample has to be diluted so that 10 L/min can be supplied to the

EEPS. For this setup, the lowest dilution ratio possible is ~100:1. If a more powerful pump were

used, a lower dilution ratio could be achieved. However, some dilution will always be required to

prevent condensation from forming. The maximum allowable sample temperature into the EEPS

is 52°C. If the raw exhaust sample is cooled to this temperature, condensate will form and

damage the EEPS. Therefore, the exhaust sample must be diluted to lower the dew point of the

sample. Calculations were performed to assess the minimum dilution ratio that can be achieved;

these calculations can be found in Appendix 9.15. For future tests, a new EEPS setup should be

used to reduce the dilution ratio. For the steady state road load conditions of these tests, no

detectable PM was observed. It is possible that by reducing the dilution ratio (i.e. having a more

concentrated exhaust sample flowing through the EEPS) the engine’s PM emissions would be

detectable with the EEPS.

95

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100

9. Appendices

Conversion from SL/min of Hydrogen to g/s 9.1.

�̇�𝐻2 =𝑃�̇�𝑀𝐻2

𝑅𝑇

�̇�𝐻2 =

(101.325 𝑘𝑃𝑎) (375𝑆𝐿𝑚𝑖𝑛) (2.01588

𝑘𝑔𝑘𝑚𝑜𝑙

) |𝑘𝑁

𝑚2⁄

𝑘𝑃𝑎| |

𝑘𝐽𝑘𝑁 ∙ 𝑚

| |𝑚3

103𝐿| |103𝑔𝑘𝑔

| |𝑚𝑖𝑛60 𝑠|

(8.314𝑘𝐽

𝑘𝑚𝑜𝑙 ∙ 𝐾) (273.15 𝐾)

�̇�𝐻2 = 0.5621𝑔

𝑠

Equivalence Ratio Calculations 9.2.

𝐻2 +𝑎

𝜑(𝑂2 + 3.773𝑁2) → 𝑏𝐻2𝑂 + 𝑐𝑂2 + 𝑑𝑁2

Atom Balance:

𝐻: 2 = 2𝑏 → 𝑏 = 1

𝑂: 2𝑎 = 𝑏 → 𝑎 = 0.5

𝑂: 2𝑎

𝜑= 𝑏 + 2𝑐 → 𝑐 =

0.5

𝜑− 0.5

𝑁: 7.546𝑎

𝜑= 2𝑑 → 𝑑 =

1.8865

𝜑

𝜒𝐻2𝑂 =𝑏

𝑏 + 𝑐 + 𝑑=

1

1 +0.5𝜑 − 0.5 +

1.8865𝜑

=1

2.3865𝜑 + 0.5

101

2.3865𝜒𝐻2𝑂

𝜑+ 0.5𝜒𝐻2𝑂 = 1 →

2.3865𝜒𝐻2𝑂

𝜑= 1 − 0.5𝜒𝐻2𝑂

𝜑 =2.3865𝜒𝐻2𝑂

1 − 0.5𝜒𝐻2𝑂

The Emissions Bench measures oxygen dry, so water is not included in the molar

balance:

𝜒𝑂2 =𝑐

𝑐 + 𝑑=

0.5𝜑 − 0.5

0.5𝜑 − 0.5 +

1.8865𝜑

=

0.5𝜑 − 0.5

2.3865𝜑 − 0.5

2.3865𝜒𝑂2𝜑

− 0.5𝜒𝑂2 =0.5

𝜑− 0.5 →

2.3865𝜒𝑂2𝜑

−0.5

𝜑= 0.5𝜒𝑂2 − 0.5

1

𝜑(2.3865𝜒𝑂2 − 0.5) = 0.5𝜒𝑂2 − 0.5

𝜑 =2.3865𝜒𝑂2 − 0.5

0.5𝜒𝑂2 − 0.5

Calculating the equivalence ratio from the air and fuel flow rates:

�̇�𝑎𝑖𝑟

�̇�𝐻2

=(𝑎𝜑)𝑀𝑂2 + (

3.773𝑎𝜑 )𝑀𝑁2

𝑀𝐻2

→�̇�𝑎𝑖𝑟𝑀𝐻2

�̇�𝐻2

=0.5𝑀𝑂2

𝜑+1.8865𝑀𝑁2

𝜑

�̇�𝑎𝑖𝑟𝑀𝐻2

�̇�𝐻2

=1

𝜑(0.5𝑀𝑂2 + 1.8865𝑀𝑁2)

𝜑 =�̇�𝐻2(0.5𝑀𝑂2 + 1.8865𝑀𝑁2)

�̇�𝑎𝑖𝑟𝑀𝐻2

102

Lubricating Oil Consumption Rate Calculations 9.3.

Calculating the mass flow rate of air going through the engine:

�̇�𝑖𝑛𝑡𝑎𝑘𝑒 =𝑉𝑑𝑁

2

𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒 = 𝑀𝑊𝑂2𝜒𝑂2 +𝑀𝑊𝑁2𝜒𝑁2 +𝑀𝑊𝐻2𝜒𝐻2

𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒 = 𝑀𝑊𝑂2 [(0.5𝜑 )

1 + (2.3865𝜑 )

] + 𝑀𝑊𝑁2 [(1.8865𝜑 )

1 + (2.3865𝜑 )

] +𝑀𝑊𝐻2 [1

1 + (2.3865𝜑 )

]

�̇�𝑒𝑛𝑔𝑖𝑛𝑒 =𝑃𝑖𝑛𝑡𝑎𝑘𝑒�̇�𝑖𝑛𝑡𝑎𝑘𝑒𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒

𝑅𝑢𝑇𝑖𝑛𝑡𝑎𝑘𝑒=𝑃𝑖𝑛𝑡𝑎𝑘𝑒𝑉𝑑𝑁𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒

2𝑅𝑢𝑇𝑖𝑛𝑡𝑎𝑘𝑒

�̇�𝑒𝑛𝑔𝑖𝑛𝑒 = �̇�𝑎𝑖𝑟 + �̇�𝐻2

(𝐴

𝐹) =


�̇�𝐻2

=(0.5𝜑 ) (𝑀𝑊𝑂2 + 3.773𝑀𝑊𝑁2)

𝑀𝑊𝐻2

�̇�𝐻2 =�̇�𝑎𝑖𝑟𝑀𝑊𝐻2

(0.5𝜑 ) (𝑀𝑊𝑂2 + 3.773𝑀𝑊𝑁2)

�̇�𝑎𝑖𝑟 =�̇�𝑒𝑛𝑔𝑖𝑛𝑒

{1 + [𝑀𝑊𝐻2

(0.5𝜑 ) (𝑀𝑊𝑂2 + 3.773𝑀𝑊𝑁2)

]}

�̇�𝑎𝑖𝑟 =(𝑃𝑖𝑛𝑡𝑎𝑘𝑒𝑉𝑑𝑁𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒

2𝑅𝑢𝑇𝑖𝑛𝑡𝑎𝑘𝑒)

{1 + [𝑀𝑊𝐻2

(0.5𝜑 ) (𝑀𝑊𝑂2 + 3.773𝑀𝑊𝑁2)

]}

103


= {

𝑃𝑖𝑛𝑡𝑎𝑘𝑒𝑉𝑑𝑁{𝑀𝑊𝑂2 [

(0.5𝜑 )

1 + (2.3865𝜑 )

] + 𝑀𝑊𝑁2 [(1.8865𝜑 )

1 + (2.3865𝜑 )

] +𝑀𝑊𝐻2 [1

1 + (2.3865𝜑 )

]}

2𝑅𝑢𝑇𝑖𝑛𝑡𝑎𝑘𝑒

}

{1 + [𝑀𝑊𝐻2

(0.5𝜑 ) (𝑀𝑊𝑂2 + 3.773𝑀𝑊𝑁2)

]}

Mass flow rate of CO2 into engine:

�̇�𝐶𝑂2,𝑖𝑛𝑡𝑎𝑘𝑒 =𝜒𝐶𝑂2,𝑏𝑎𝑐𝑘𝑔𝑟𝑜𝑢𝑛𝑑𝑀𝑊𝐶𝑂2�̇�𝑎𝑖𝑟

𝑀𝑊𝑎𝑖𝑟

Mass flow rate of CO2 out of engine:

�̇�𝐶𝑂2,𝑒𝑥ℎ𝑎𝑢𝑠𝑡 =𝜒𝐶𝑂2𝑀𝑊𝐶𝑂2�̇�𝑒𝑛𝑔𝑖𝑛𝑒

𝑀𝑊𝑒𝑥ℎ𝑎𝑢𝑠𝑡

Mass flow rate of CO2 produced by the engine:

�̇�𝐶𝑂2,𝑝𝑟𝑜𝑑𝑢𝑐𝑒𝑑 = �̇�𝐶𝑂2,𝑒𝑥ℎ𝑎𝑢𝑠𝑡 − �̇�𝐶𝑂2,𝑖𝑛𝑡𝑎𝑘𝑒

Mass Flow rate of CO produced by the engine:

�̇�𝐶𝑂 =𝜒𝐶𝑂𝑀𝑊𝐶𝑂�̇�𝑒𝑛𝑔𝑖𝑛𝑒


104

Mass flow rate of lubricating oil consumed:

�̇�𝑜𝑖𝑙 =�̇�𝐶𝑂2,𝑝𝑟𝑜𝑑𝑢𝑐𝑒𝑑 [𝑀𝑊𝑐+𝑀𝑊𝐻 (

𝐻𝐶)]

𝑀𝑊𝐶𝑂2

+�̇�𝐶𝑂 [𝑀𝑊𝑐+𝑀𝑊𝐻 (

𝐻𝐶)]

𝑀𝑊𝐶𝑂

It is assumed that the H/C ratio is 0.86.

Converting Emissions to a Per km Basis 9.4.

Emissions measured in volumetric concentrations can be converted to a mass basis per

kilometer with the following analysis. The mass of fluid going through the engine and molecular

weight of the exhaust is calculated in the same was as described in Appendix 9.3.

𝑀𝑊𝑁𝑂𝑥 = 30.01 ×%𝑁𝑂 + 46.055 ×%𝑁𝑂2

�̇�𝑁𝑂𝑥 = 𝜒𝑁𝑂𝑥�̇�𝑒𝑛𝑔𝑖𝑛𝑒

𝑀𝑊𝑁𝑂𝑥


𝑚𝑁𝑂𝑥 =�̇�𝑁𝑂𝑥

𝑆𝑝𝑒𝑒𝑑 𝑜𝑓 𝑉𝑒ℎ𝑖𝑐𝑙𝑒=

�̇�𝑁𝑂𝑥

100𝑘𝑚ℎ𝑟

LICOR CO2 Measurement Correction 9.5.

The Licor measures after a heat exchanger which removes most of the water. Therefore,

the CO2 measurement needs to be corrected for the missing water. If you assume that almost no

CO2 is produced by the engine, the following analysis holds:

105

𝐻2 +𝑎

𝜑[𝑂2 + (3.773 − 4.773𝜒𝐶𝑂2,𝑏𝑎𝑐𝑘𝑔𝑟𝑜𝑢𝑛𝑑)𝑁2 + 4.773𝜒𝐶𝑂2,𝑏𝑎𝑐𝑘𝑔𝑟𝑜𝑢𝑛𝑑𝐶𝑂2]

→ 𝑏𝐻2𝑂 + 𝑐𝑂2 + 𝑑𝑁2 + 𝑒𝐶𝑂2

𝑎 = 0.5

𝑏 = 1

𝑐 =0.5

𝜑− 0.5

𝑑 =3.773 − 4.773𝜒𝐶𝑂2,𝑏𝑎𝑐𝑘𝑔𝑟𝑜𝑢𝑛𝑑

2𝜑

𝑒 =2.3865𝜒𝐶𝑂2,𝑏𝑎𝑐𝑘𝑔𝑟𝑜𝑢𝑛𝑑

𝜑

𝜒𝐶𝑂2,𝑙𝑖𝑐𝑜𝑟 =𝑒

𝑏𝑙𝑜𝑤 + 𝑐 + 𝑑 + 𝑒

𝜒𝐶𝑂2,𝑒𝑥ℎ𝑎𝑢𝑠𝑡 =𝑒

𝑏 + 𝑐 + 𝑑 + 𝑒

𝑒 =𝜒𝐶𝑂2,𝑙𝑖𝑐𝑜𝑟(𝑏𝑙𝑜𝑤 + 𝑐 + 𝑑)

1 − 𝜒𝐶𝑂2,𝑙𝑖𝑐𝑜𝑟

𝜒𝐻2𝑂,𝑙𝑖𝑐𝑜𝑟 =𝑏𝑙𝑜𝑤

𝑏𝑙𝑜𝑤 + 𝑐 + 𝑑 + 𝑒

𝑏𝑙𝑜𝑤 =𝜒𝐻2𝑂,𝑙𝑖𝑐𝑜𝑟(𝑐 + 𝑑 + 𝑒)

1 − 𝜒𝐻2𝑂,𝑙𝑖𝑐𝑜𝑟

106

𝜒𝐶𝑂2,𝑒𝑥ℎ𝑎𝑢𝑠𝑡 ={

𝜒𝐶𝑂2,𝑙𝑖𝑐𝑜𝑟 {[

𝜒𝐻2𝑂,𝑙𝑖𝑐𝑜𝑟 (0.5𝜑− 0.5 +

3.773 − 4.773𝜒𝐶𝑂2,𝑏𝑎𝑐𝑘𝑔𝑟𝑜𝑢𝑛𝑑2𝜑

+2.3865𝜒𝐶𝑂2,𝑏𝑎𝑐𝑘𝑔𝑟𝑜𝑢𝑛𝑑

𝜑)

1 − 𝜒𝐻2𝑂,𝑙𝑖𝑐𝑜𝑟] +

0.5𝜑− 0.5 +


}

1 − 𝜒𝐶𝑂2,𝑙𝑖𝑐𝑜𝑟

}

(0.5𝜑+ 0.5 +


+2.3865𝜒𝐶𝑂2,𝑏𝑎𝑐𝑘𝑔𝑟𝑜𝑢𝑛𝑑

𝜑)

Fuel Contribution to CO and CO2 9.6.

One important question is whether the CO in the exhaust could be from the hydrogen

tank instead of the lubricating oil combustion. This analysis will try to answer that question. The

hydrogen tanks did not specify the CO level, but they did specify that the hydrogen content was

more than 99.998% of the cylinder’s volume. Therefore, the following analysis is a worst case

scenario:

𝜒𝐶𝑂,𝑓𝑢𝑒𝑙 = 0.00002

The 0.6 equivalence ratio would be the worst scenario for this analysis, so it will be used

for all of the calculations. In the intake, the molar ratios are:

𝜒𝐻2 =1

1 +2.3865𝜑

=1

1 +2.38650.6

= 0.200904068

𝜒𝑂2 =

0.5𝜑

1 +2.3865𝜑

=

0.50.6

1 +2.38650.6

= 0.167420056

𝜒𝑁2 =

1.8865𝜑

1 +2.3865𝜑

=

1.88650.6

1 +2.38650.6

= 0.631675874

𝜒𝐶𝑂,𝑖𝑛𝑡𝑎𝑘𝑒 = 𝜒𝐻2𝜒𝐶𝑂,𝑓𝑢𝑒𝑙 = (0.200904068)(0.00002) = 4.01808136 × 10−6

107

𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒 = 𝑀𝑊𝑂2𝜒𝑂2 +𝑀𝑊𝑁2𝜒𝑁2 +𝑀𝑊𝐻2𝜒𝐻2

𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒 = (31.9988𝑘𝑔

𝑘𝑚𝑜𝑙) (0.167420056) + (28.0134

𝑘𝑔

𝑘𝑚𝑜𝑙) (0.631675874)

+ (2.01588𝑘𝑔

𝑘𝑚𝑜𝑙) (0.200904068)

𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒 = 23.45762831𝑘𝑔

𝑘𝑚𝑜𝑙

In the exhaust, the molar ratios are:

𝜒𝐻2𝑂 =1

2.3865𝜑 + 0.5

=1

2.38650.6 + 0.5

= 0.223338916

𝜒𝑂2 =

0.5𝜑 − 0.5

2.3865𝜑

+ 0.5=

0.50.6 − 0.5

2.38650.6

+ 0.5= 0.074446305

𝜒𝑁2 =

1.8865𝜑

2.3865𝜑 + 0.5

=

1.88650.6

2.38650.6 + 0.5

= 0.702214777

𝑀𝑊𝑒𝑥ℎ𝑎𝑢𝑠𝑡 = 𝑀𝑊𝑂2𝜒𝑂2 +𝑀𝑊𝑁2𝜒𝑁2 +𝑀𝑊𝐻2𝑂𝜒𝐻2𝑂

𝑀𝑊𝑒𝑥ℎ𝑎𝑢𝑠𝑡 = (31.9988𝑘𝑔

𝑘𝑚𝑜𝑙) (0.074446305) + (28.0134

𝑘𝑔

𝑘𝑚𝑜𝑙) (0.702214777)

+ (18.01528𝑘𝑔

𝑘𝑚𝑜𝑙) (0.223338916)

𝑀𝑊𝑒𝑥ℎ𝑎𝑢𝑠𝑡 = 26.07712897𝑘𝑔

𝑘𝑚𝑜𝑙

108

𝑚𝑖𝑛𝑡𝑎𝑘𝑒 = 𝑚𝑒𝑥ℎ𝑎𝑢𝑠𝑡

𝑚𝐶𝑂,𝑖𝑛𝑡𝑎𝑘𝑒 = 𝑚𝐶𝑂,𝑒𝑥ℎ𝑎𝑢𝑠𝑡

𝜒𝐶𝑂,𝑖𝑛𝑡𝑎𝑘𝑒𝑀𝑊𝐶𝑂

𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒=𝜒𝐶𝑂,𝑒𝑥ℎ𝑎𝑢𝑠𝑡𝑀𝑊𝐶𝑂


𝜒𝐶𝑂,𝑒𝑥ℎ𝑎𝑢𝑠𝑡 = (𝑀𝑊𝑒𝑥ℎ𝑎𝑢𝑠𝑡

𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒) 𝜒𝐶𝑂,𝑖𝑛𝑡𝑎𝑘𝑒

𝜒𝐶𝑂,𝑒𝑥ℎ𝑎𝑢𝑠𝑡 = (26.07712897


23.45762831𝑘𝑔𝑘𝑚𝑜𝑙

) (4.01808136 × 10−6) = 4.466778331 × 10−6

≅ 4.47 𝑝𝑝𝑚

The highest CO concentration observed at the road load condition was ~4.75 ppm. So, in

general, it is possible that the CO concentration increase observed was almost solely from the

hydrogen tanks. However, it is unlikely that the CO levels were this high and that the CO being

injected into the cylinder before combustion would survive. If CO was injected alongside the

fuel, it would be oxidized in the combustion chamber into CO2.

Available Turbocharger Power 9.7.

The power available to the turbocharger can be calculated by assuming a constant exit

temperature of the exhaust side.

𝑃𝑡𝑢𝑟𝑏𝑜𝑐ℎ𝑎𝑟𝑔𝑒𝑟 = �̇�𝑒𝑥ℎ𝑎𝑢𝑠𝑡𝑐𝑝(𝑇𝑒𝑥ℎ𝑎𝑢𝑠𝑡 − 300°𝐶)

For simplicity, the value for 𝑐𝑝 was assumed to be a constant 1.1𝑘𝐽

𝑘𝑔∙𝐾

109

Road Load Power 9.8.

An engine condition (speed and torque) needs to be selected for the tests. The condition

for these experiments will be determined based on the power required to drive a 2001 Ford

Ranger on a level road at 100 km/hr. The data used in this section was taken from Alin Pop’s

master’s thesis (Pop, 2016). The road load power of the engine is given by the following

equation:

𝑃𝑟 = (𝐶𝑅 ∙ 𝑀𝑣 ∙ 𝑔 +1

2𝜌𝑎𝑖𝑟 ∙ 𝐶𝐷 ∙ 𝐴𝑣 ∙ 𝑆𝑣

2) 𝑆𝑣

𝑃𝑟 = (2.73𝐶𝑅 ∙ 𝑀𝑣(𝑘𝑔) + 0.0126𝐶𝐷 ∙ 𝐴𝑣(𝑚2) ∙ 𝑆𝑣 (

𝑘𝑚

ℎ𝑟)2

) 𝑆𝑣 (𝑘𝑚

ℎ𝑟) × 10−3

Table 5. Variables for Road Load Power Calculation

Parameter Variable Imperial Value Metric Value

Coefficient of Rolling

Resistance 𝐶𝑅 0.0135 0.0135

Curb Weight 𝑀𝑣 ∙ 𝑔 3100 lbs 13789 N

Coefficient of

Aerodynamic Drag 𝐶𝐷 0.49 0.49

Vehicle Frontal Area 𝐴𝑣 25.9 ft² 2.41 m²

Gear Ratios

1st 2.47:1 2.47:1

2nd

1.87:1 1.87:1

3rd

1.47:1 1.47:1

4th

1.00:1 1.00:1

5th

0.75:1 0.75:1

Final Drive 3.73:1 3.73:1

𝑃𝑟 = [(2.73)(0.0135)(1406 𝑘𝑔) + (0.0126)(0.49)(2.41 𝑚2) (100𝑘𝑚

ℎ𝑟)2

] (100𝑘𝑚

ℎ𝑟) × 10−3

𝑃𝑟 = 20.06 𝑘𝑊

110

The angular velocity of the tires at 100 km/hr is:

𝜔𝑡𝑖𝑟𝑒 = (496𝑟𝑒𝑣

𝑘𝑚)(100

𝑘𝑚

ℎ𝑟) (

ℎ𝑟

60 𝑚𝑖𝑛) = 827

𝑟𝑒𝑣

𝑚𝑖𝑛

Assuming that the vehicle is in fifth gear, the speed of the engine is:

𝜔𝑒𝑛𝑔𝑖𝑛𝑒 = 𝐺𝑅𝑓𝑖𝑛𝑎𝑙𝐺𝑅5𝑡ℎ𝜔𝑡𝑖𝑟𝑒 = (3.73)(0.75) (827𝑟𝑒𝑣

𝑚𝑖𝑛) = 2314

𝑟𝑒𝑣

𝑚𝑖𝑛

Given the speed of the engine, the torque setting can be calculated:

𝑇 =𝑃𝑟

2𝜋𝜔𝑒𝑛𝑔𝑖𝑛𝑒=(20.06 𝑘𝑊) |

103𝑊𝑘𝑊

| |𝐽/𝑠𝑊 | |

𝑁 ∙ 𝑚𝐽 | |

60 𝑠𝑚𝑖𝑛|

(2)(𝜋) (2314𝑟𝑒𝑣𝑚𝑖𝑛)

= 82.78 𝑁 ∙ 𝑚

Supercharger Power Calculations 9.9.

The mechanical power required to drive the supercharger was estimated in two ways.

First, Figure 22 shows that the average fuel conversion efficiency of the engine is 28.1% for the

supercharged tests and 35.1% for the naturally aspirated tests. If the fuel conversion efficiency of

the naturally aspirated engine, fuel flow rate of the supercharged engine, and power output of the

supercharged engine are used, the power required by the supercharger can be calculated based on

the following formula:

𝑃𝑠𝑢𝑝𝑒𝑟𝑐ℎ𝑎𝑟𝑔𝑒𝑟 = 𝜂𝑓𝑐,𝑛𝑎𝑡𝑢𝑟𝑎𝑙𝑙𝑦 𝑎𝑠𝑝𝑖𝑟𝑎𝑡𝑒𝑑�̇�𝐻2,𝑠𝑢𝑝𝑒𝑟𝑐ℎ𝑎𝑟𝑔𝑒𝑑𝐿𝐻𝑉𝐻2 − 𝑃𝑒𝑛𝑔𝑖𝑛𝑒,𝑠𝑢𝑝𝑒𝑟𝑐ℎ𝑎𝑟𝑔𝑒𝑑

111

This calculation results in a supercharger work of 4.9 kW. In addition to calculating the

parasitic power of the supercharger from the fuel conversion efficiency, the power required by

the supercharger can also be calculated with the Willan’s line.

The Willan’s line is generated by plotting the engine power generated vs. fuel flow rate at

a constant speed. By graphing the power vs. fuel flow rate, a linear relation emerges. This linear

relation can be extrapolated backwards to find the frictional power of the engine at zero fuel flow

rate. For these tests, the Willan’s lines were generated by plotting the power and fuel flow rate of

the engine during the engine loading process as the throttle plate was opened. The first data point

in each series is taken once the engine passes ~2300 RPM which is the road load speed. The last

point in the series is taken when the engine reaches the road load power condition ~83 N·m. All

nine supercharged tests and three naturally aspirated tests were graphed and then a linear best

approximation function was fit for each case, as shown in Figure 37 below.

Figure 37. Willan’s Line for Supercharged and Naturally Aspirated Tests

y = 0.0986x - 19.3416

y = 0.1059x - 13.0999

-40

-30

-20

-10

0

10

20

30

40

50

0 100 200 300 400 500 600

Engi

ne

Po

we

r (k

W)

Hydrogen Flowrate (SL/min)

Willan's Line - Engine Power (kW) vs. Hydrogen Flowrate (SL/min) for

Supercharged and Naturally Aspirated Tests

Linear (Supercharged)

Linear (Naturally Aspirated)

112

Since the fuel conversion efficiencies of all three supercharger equivalence ratios were

statistically the same, all three supercharged equivalence ratios were grouped together for the

supercharged Willan’s line. By subtracting the y-intercept of the naturally aspirated Willan’s line

from the y-intercept of the supercharged Willan’s line, the power required to operate the

supercharger can be calculated. Therefore, based on the Willan’s line analysis, the power

required to overcome the friction of the engine is 13.1 kW and the power required to operate the

supercharger is 6.2 kW.

The two methods are in reasonable agreement considering that the power to drive the

supercharger was not directly measured.

Throttle Body Arduino Code 9.10.

/*

digital pins:

2 - direction

3 - step

4 - MS2

5 - MS1

6 - MS0

7 - enable

8 to 13 - lcd

analog pins:

0 - dial on front of electrical box

1 - throttle position sensor

*/

#include <LiquidCrystal.h>

LiquidCrystal lcd(8, 9, 10, 11, 12, 13);

char user_input;

int x = 0;

void setup()

113

{

lcd.begin(20,4);

Serial.begin(9600);

Serial.println("Begin motor control");

Serial.println();

Serial.println("Enter number for control option:");

Serial.println("1. Forward One Steps");

Serial.println("2. Forward Ten Step");

Serial.println("3. Back One Steps");

Serial.println("4. Back Ten Step");

Serial.println();

digitalWrite(6, HIGH); //set the step size to thirty second

digitalWrite(5, HIGH);

digitalWrite(4, HIGH);

}

void loop()

{

while(Serial.available())

{

user_input = Serial.read(); //Read user input and trigger appropriate function

digitalWrite(7, LOW); //Pull enable pin low to set FETs active and allow motor control

if (user_input =='1')

{

OneStepForward();

}

else if(user_input =='2')

{

TenStepsForward();

}


{

OneStepBackward();

}


{

TenStepsBackward();

}

else

{

Serial.println("Invalid option entered.");

}

}

114

}

void OneStepForward()

{

Serial.println("Moving One Forward");

digitalWrite(2, HIGH); //Pull direction pin low to move "forward"

for(x= 0; x<1; x++) //Loop the forward stepping enough times for motion to be visible

{

digitalWrite(3,HIGH); //Trigger one step forward

delay(1);

digitalWrite(3,LOW); //Pull step pin low so it can be triggered again

delay(1);

}

Serial.println("Enter New Command");

Serial.println();

}

void TenStepsForward()

{

Serial.println("Moving Ten Forward");

digitalWrite(2, HIGH); //Pull direction pin low to move "forward"


{

digitalWrite(3,HIGH); //Trigger one step forward

delay(1);


delay(1);

}


Serial.println();

}

void OneStepBackward()

{

Serial.println("Moving One Backward");

digitalWrite(2, LOW); //Pull direction pin low to move "backward"


{

digitalWrite(3,HIGH); //Trigger one step backward

delay(1);


delay(1);

}


Serial.println();

}

115

void TenStepsBackward()

{

Serial.println("Moving Ten Backward");

digitalWrite(2, LOW); //Pull direction pin low to move "backward"


{

digitalWrite(3,HIGH); //Trigger one step backward

delay(1);


delay(1);

}


Serial.println();

}

Clean Room Procedure 9.11.

1) Ensure temperature is 71.6°F ± 1.8°F and humidity is 45.0 % ± 5.0 %

2) Turn on light to clean room

3) Clean shoes on sticky mat

4) Enter clean room

5) Put on shoe covers

6) Put on gloves

7) Put on hood

8) Put on suit (hold sleeves while putting on so that they don’t touch the ground)

9) Button up wrists and ankles

10) Wipe down Tupperware if you’re bringing anything in to the clean room

11) Enter clean room

12) Wipe down front of desk, computer, and tweezers with wet wipe

13) Turn on scale and red irradiating light

14) Turn on computer

15) Open recording sheet and online timer

16) Once the scale reads “cal” open scale lid and wait 20 seconds

17) Close scale lid and wait until “mg” is on screen and then wait 10 seconds

18) Obtain filter to weigh

116

19) Once the scale is mg for 10 seconds put the filter under the red light for 10 seconds on

each side

20) Open the scale lid and place the filter on the scale

21) Waits 20 seconds

22) Close the lid of the scale

23) Start 2 minutes, 30 seconds on the online timer

a. While the timing is going take a filter holder out, clean it and write its number on

the filter’s old clear holder

24) Record weight at the end of the time

25) Open the scale lid

26) Place the filter directly into its filter holder and place the filter holder into the

Tupperware

27) Repeat until all of the filters are done

28) Turn off the scale and the red irradiating light

29) Take any garbage and the Tupperware

30) Close the computer program

31) Exit the clean room and take off the outfit in the reverse order with which you put it on

32) Write down the temperature and relative humidity in the book outside the clean room

Emissions Operating Procedure 9.12.

Emissions Bench

33) Turn on emissions bench with 2 power bars inside the cabinet at the bottom (30 minutes

prior to calibration)

34) Turn on the helifuel and set it to 15 psig

35) Turn on the zero air and set it to 15 psig

36) Turn on the pump (green rocker switch on the bottom of the front cabinet)

37) Press main and then F1 to switch to measurement mode on the bottom three boxes

38) Press main and then F8 to get ignition on the top box

a. Wait 20 seconds

117

39) Press F3 on the top box to check the flame temperature (should be ~300°C)

40) Leave the emissions bench to warm up for 30 minutes

41) Turn on N2 tank and set to 10 psig

42) Press main on all four boxes

43) Make sure that the valve on the exhaust is sampling from the span gas channel and not

the exhaust stream

44) On flow control box, flip the right hand rocker switch to manual and turn the knob to zero

45) On the back of the emissions bench, turn the span/zero valve from room air to cal gas

46) Move quickly to the front of the emissions bench and turn the far right rotameter until the

gauge reads 10 psig

a. Wait for a minute for the flow to stabilize

47) Press main, F4, F2, F1, F1 for all of the boxes

48) Turn on THC (propane) tank and set to 10 psig

49) Turn off the nitrogen tank

50) Go quickly to the front of the emissions bench and when the gauge starts to drop, switch

the knob from zero to number one and turn the rotameter knob until the pressure

stabilizes at 10 psig

51) On the top box, F3 (range select), select the appropriate range for the tank, press manual

cal, and then when the value stabilizes press save

52) Switch the span/zero valve on the back of the emissions bench to room air

53) Turn off the THC tank

54) Switch the THC valve on the back of the emissions bench to trench and then back to cal

gas after a couple of seconds

55) Take off the regulator from the THC tank and put it on the new tank (be careful because

it’s a left-hand thread)

56) Turn on the new THC tank and set to 10 psig

57) Switch span/zero valve to cal gas (on the back of the emissions bench)

58) Go quickly to the front of the emissions bench and turn the rotameter until the gauge is

10 psig

59) Go through the same procedure to calibrate NOx

60) Close NOx tank

118

61) On back of emissions bench turn span/zero valve to room air

62) Take off NOx regulator and put it onto the new NOx tank and set to 10 psig

63) Turn valve on back of emissions bench to cal gas

64) Turn rotameter until gauge reads 10 psig

65) Then same procedure for new range

66) Then same procedure for third NOx span gas

67) Open CO2 tank and set to 10 psig

68) On third box from the top, switch setting to CO2 from O2 (CO2 is on left and O2 is on

right)

69) Turn span/zero valve to cal gas and switch knob to gas 3 on front of emissions bench

70) Go through the same calibration procedure for CO2

71) Turn on O2 tank and turn off CO2 tank, switch to gas 4 when gauge starts to drop

72) Go through the same calibration procedure for O2

73) Turn off the O2 tank and turn on the CO tank

74) When the gauge drops switch to span 5

75) Go through the same calibration procedure for CO

76) Turn off the CO tank and when the gauge drops switch to room air on span/zero valve

77) Set knob to off on flow box and switch far right rocker switch from “manual to auto”

FTIR

78) Fill the dewar with liquid nitrogen (full 4 litres) and turn valve from ‘run’ to ‘vent’

EEPS

79) Open the orange thermos diluter backing (four long screws)

80) Wipe the inside with a chemwipe to dry (also wipe the inside of the disk)

81) Ensure that the disk is sitting flush against the matting face when it is reinstalled.

EEPS

82) Ensure that the black valve on top of the EEPS is closed

83) Turn on the wall supply of air (red valve) and set the regulator to 15 psig (where the clear

tube tees off)

119

FTIR

84) Put the cap on the dewar and turn the valve under the dewar towards ‘run’ when the valve

is completely frosted over and there is a steady stream of N2 gas flowing out

EEPS

85) Turn the top box on (the rocker switch is located on the back of the machine next to the

power cord)

86) From left to right on the front panel of the top box

a. Pump switch up

b. Temperature knob from OFF to 80°C

c. Dil air up

d. Air supply up

e. On the far right hand side of the top box where the controller is

i. Mode

ii. Up until 300

iii. Mode again to set

iv. Thc switch up

FTIR

87) FTIR config utilities – instrument monitor

88) Write today’s date in the book

89) Max signal (read off of graph)

90) Average between 100 and 500

91) Phase angle (bottom right)

92) Phi pp (bottom right)

93) igram max (bottom left)

94) igram min (bottom left)

95) DC level (bottom left)

96) FTIR diagnostics

a. SNR test

b. Config FTIR

120

c. Set number of scans to 20

d. Press okay

e. Run SNR test

f. Write down SNR tests

g. Check that all of the things written down are within range of the sheet values

h. Return

97) Multigas main

a. Setup

b. Select file

c. Make a folder for the file called today’s date

d. Inside that folder enter the file name as today’s date and press okay

e. Set directory

f. Find the folder you just made

g. Press ‘select current directory’

h. Basename – make it the date

i. Press okay in bottom right

EEPS

98) Take off the filter on the back

99) Hook up the output of the thermos diluter to the back of the EEPS where the filter was

100) Turn on the valve on top of the EEPS until the flow meter goes from a large value

to zero and then back to 3 (sucking to blowing)

a. Halfway when you get to blowing 1, take off the blue filter

b. Hook up the exhaust tube to the back of the EEPS that goes to the trench

c. Turn on fan to cool the pump drawing the sample out of the exhaust

LICOR

101) In Li840a program

102) Connect symbol

103) Red log button

104) Set name to today’s date

121

FTIR

105) Close red knob on rotameter on FTIR Purge Gas generator

106) Check that rotameters on the front are above 4

107) Set purge gas between 6 and 7 psig (front of FTIR)

108) Multigas main > run > new background (wait) > return

109) Open red rotameter knob back up (quickly)

110) Adjust front pressure gauge back to 6 or 7 psig

111) Adjust rotameters on front above 4

EEPS

112) Turn on EEPS Labview file on computer

a. New file

b. Today’s date

c. Open

d. Record for 10 minutes before running

Emissions Bench

113) Turn valve on exhaust to sample exhaust gas

FTIR

114) Run in program

115) Push in stick on back

116) Zero off (switch)

117) Pump on (rocker switch)

Allow for a ten minute background of all of the emissions equipment

Turn on Engine

122

After Test Run

EEPS

118) Stop logging

FTIR

119) Pump off

120) Zero on

121) Pull plunger out

122) Press return on computer program

EEPS

123) Mode

124) Down until 150°C

125) On computer > go to file > export to save the file

Emissions Bench

126) Switch from sampling exhaust to room air

127) Turn off helifuel supply

128) Turn off zero air supply

EEPS

129) Once at 150°C

a. Thc off

b. Air supply off

c. Dil air off

d. Temp from 80°C to off

e. Pump off

f. Back power bar off

g. Turn black valve on top of EEPS until closed

h. Close wall air supply (red valve on wall)

123

i. Unplug flow meter

j. Remove inlet and outlet tubes (black)

k. Replace inlet tube with filter

Emissions Bench

130) Main, F7 for all four screens to set to standby

131) Leave pump on for 30 minutes to clear all of the water

132) Relieve helifuel through PRV

133) Turn off pump (green rocker switch on bottom of front)

124

Calibration Procedure for Sensors 9.13.

9.13.1. Mass Air Flow Sensor

Nomenclature

𝑑 ≈ diameter of top moving cylinder in bell prover

𝑙 ≈ length of travel of top moving cylinder in bell prover

∆𝑡 ≈ time between actuation of micro switches

𝜌 ≈ density of air

𝑉 ≈ volume of top moving cylinder in bell prover

�̇� ≈ volumetric flow rate of air through MAF sensor

�̇� ≈ mass flow rate of air through MAF sensor

125

Equipment Setup

Figure 38. MAF Sensor Calibration Configuration with Bell Prover

Procedure

1. The Mass Air Flow (MAF) sensor was connected to a car battery; pin 1 to positive and

pin 2 to negative in that order.

2. The MAF sensor was connected to a multimeter to measure the sensor’s analog output.

Pin 3 was connected to common and pin 4 was connected to the voltage input.

3. The ball valve on the bell prover was opened.

4. The top moving cylinder in the bell prover was lowered so that its top was below the first

micro switch.

5. The timing program was loaded onto the Arduino which was connected to the two micro

switches.

126

6. The globe valve which throttled the compressed air from the wall supply was slowly

opened to reach each one of the desired voltage data points from the multimeter.

7. Once the desired set point was reached on the multimeter by throttling the compressed air

flow with the globe valve, the flow was left for a couple of minutes to stabilize.

8. The ball valve on the bell prover was closed. This action caused the air to accumulate in

the bell prover which caused the top moving cylinder in the bell prover to rise and pass

the first micro switch.

9. When the top moving cylinder in the bell prover reached the end of its scale and passed

the second micro switch, the ball valve on the bell prover was opened.

10. The time delay between the actuation of the two micro switches, which was displayed by

the Arduino, was recorded.

11. Each desired voltage set point was repeated three times for accuracy.

12. Steps 7 to 11 were repeated for the entire range of desired voltage set points.

127

Results


Test on July 31st, 2014

Test Δt (ms) E

(V) Test

Δt

(ms)

E

(V) Test

Δt

(ms)

E

(V) Test

Δt

(ms)

E

(V) Test

Δt

(ms)

E

(V)

0 N/A 0.04 10 35020 1.25 20 13072 2.04 30 6280 2.76 40 2995 3.71

1 139185 0.51 11 35440 1.25 21 13102 2.03 31 5146 3.02 41 3025 3.70

2 138113 0.52 12 35424 1.25 22 10395 2.24 32 5135 3.01 42 2867 3.78

3 140366 0.50 13 24071 1.53 23 10396 2.25 33 5233 3.00

4 81816 0.76 14 24280 1.52 24 10454 2.24 34 4265 3.25

5 83974 0.76 15 24267 1.53 25 7786 2.53 35 4336 3.23

6 79329 0.78 16 17755 1.76 26 7813 2.53 36 4316 3.23

7 46979 1.06 17 17870 1.75 27 7855 2.52 37 3522 3.49

8 48088 1.04 18 17831 1.76 28 6343 2.76 38 3551 3.48

9 48927 1.05 19 12946 2.05 29 6298 2.77 39 3635 3.44


Test on August 11th

, 2014

Test Δt (ms) E (V) Test Δt

(ms)

E

(V) Test

Δt

(ms)

E

(V) Test

Δt

(ms)

E

(V) Test

Δt

(ms)

E

(V)

43 N/A 0.04 54 33299 1.29 65 9844 2.29 76 5257 2.97 87 2520 3.95

44 148277 0.47 55 33640 1.28 66 9846 2.29 77 4249 3.23 88 2549 3.93

45 136535 0.51 56 25520 1.49 67 10065 2.28 78 4234 3.24 89 2160 4.19

46 143418 0.49 57 25610 1.48 68 7951 2.50 79 4311 3.22 90 2324 4.09

47 79514 0.77 58 25942 1.46 69 7950 2.50 80 3633 3.44 91 2111 4.25

48 81598 0.76 59 17552 1.77 70 7949 2.50 81 3730 3.41 92 2117 4.24

49 83297 0.76 60 17905 1.77 71 6349 2.75 82 3794 3.38 93 2067 4.27

50 50689 1.01 61 17905 1.77 72 6440 2.73 83 2769 3.81 94 2126 4.24

51 48636 1.03 62 13133 2.03 73 6476 2.73 84 2715 3.84

52 49504 1.02 63 13194 2.03 74 5195 2.98 85 2862 3.76

53 32960 1.30 64 13317 2.02 75 5279 2.98 86 2465 3.99

128

Calculations

Table 8. MAF Sensor Output Voltage and Calculated Mass Air Flow Rate from Tests on July

31st, 2014 and August 11

th, 2014

Test E

(V)

ṁ

(kg/hr) Test

E

(V)

ṁ

(kg/hr) Test

E

(V)

ṁ

(kg/hr) Test

E

(V)

ṁ

(kg/hr) Test

E

(V)

ṁ

(kg/hr)

0 0.04 0.00 20 2.04 59.88 40 3.71 261.34 60 1.77 43.71 80 3.44 215.45

1 0.51 5.62 21 2.03 59.74 41 3.70 258.75 61 1.77 43.71 81 3.41 209.84

2 0.52 5.67 22 2.24 75.30 42 3.78 273.01 62 2.03 59.60 82 3.38 206.30

3 0.50 5.58 23 2.25 75.29 43 0.04 0.00 63 2.03 59.32 83 3.81 282.67

4 0.76 9.57 24 2.24 74.87 44 0.47 5.28 64 2.02 58.78 84 3.84 288.29

5 0.76 9.32 25 2.53 100.53 45 0.51 5.73 65 2.29 79.51 85 3.76 273.48

6 0.78 9.87 26 2.53 100.18 46 0.49 5.46 66 2.29 79.50 86 3.99 317.53

7 1.06 16.66 27 2.52 99.65 47 0.77 9.84 67 2.28 77.77 87 3.95 310.60

8 1.04 16.28 28 2.76 123.40 48 0.76 9.59 68 2.50 98.44 88 3.93 307.07

9 1.05 16.00 29 2.77 124.28 49 0.76 9.40 69 2.50 98.45 89 4.19 362.37

10 1.25 22.35 30 2.76 124.64 50 1.01 15.44 70 2.50 98.47 90 4.09 336.80

11 1.25 22.09 31 3.02 152.10 51 1.03 16.09 71 2.75 123.28 91 4.25 370.78

12 1.25 22.10 32 3.01 152.43 52 1.02 15.81 72 2.73 121.54 92 4.24 369.73

13 1.53 32.52 33 3.00 149.57 53 1.30 23.75 73 2.73 120.86 93 4.27 378.67

14 1.52 32.24 34 3.25 183.52 54 1.29 23.51 74 2.98 150.67 94 4.24 368.16

15 1.53 32.25 35 3.23 180.52 55 1.28 23.27 75 2.98 148.27

16 1.76 44.08 36 3.23 181.35 56 1.49 30.67 76 2.97 148.89

17 1.75 43.80 37 3.49 222.24 57 1.48 30.56 77 3.23 184.21

18 1.76 43.90 38 3.48 220.42 58 1.46 30.17 78 3.24 184.86

19 2.05 60.46 39 3.44 215.33 59 1.77 44.59 79 3.22 181.56

129

Sample Calculation for Test #1

𝑑 = 26.51 𝑖𝑛 ± 0.06 𝑖𝑛

𝑙 = 20.03 𝑖𝑛 ± 0.01 𝑖𝑛

∆𝑡 = 139185 𝑚𝑠 ± 0.0057 𝑚𝑠

𝜌 = 1.20𝑘𝑔

𝑚3 ± 0.02

𝑘𝑔

𝑚3

𝑉 = (𝜋) (𝑑

2)2

(𝑙) = (𝜋) (26.51 𝑖𝑛

2)2

(20.03 𝑖𝑛) |𝑚3

61023.7 𝑖𝑛3| = 0.1809 𝑚3

�̇� =𝑉

∆𝑡= (

0.1809 𝑚3

139185 𝑚𝑠) |3.6 × 106 𝑚𝑠

ℎ𝑟| = 4.6790

𝑚3

ℎ𝑟

�̇� = 𝜌�̇� = (1.20𝑘𝑔

𝑚3)(4.6790

𝑚3

ℎ𝑟) = 5.61

𝑘𝑔

ℎ𝑟

Figure 39. Graph of Mass Air Flow Rate (kg/hr) vs. MAF Sensor Output Voltage (V)

y = 3.6731x3 + 3.3761x2 + 7.4834x + 0.3205

0.00

100.00

200.00

300.00

400.00

0.00 0.50 1.00 1.50 2.00 2.50 3.00 3.50 4.00 4.50

Mas

s A

ir F

low

Rat

e (

kg/h

r)

MAF Sensor Output Voltage (V)

Mass Air Flow Rate (kg/hr) vs. MAF Sensor Output Voltage (V)

130

Arduino Code

unsigned long time_one=0;

unsigned long time_two=0;

unsigned long n=0;

int test=0;

void setup()

{

Serial.begin(9600);

}

void loop()

{

n=0;

if(digitalRead(2)==HIGH)

{

time_one=millis();

while(digitalRead(3)==LOW)

{

n=n+1;

}

time_two=millis();

test=test+1;

double nu=(double)n;

double diff=(double)(time_two-time_one);

Serial.print("Test "); Serial.print(test); Serial.print(": ");

Serial.print(time_two-time_one); Serial.print(" milliseconds +- ");

Serial.print((diff/nu)*1000); Serial.println(" microseconds");

}

}

131

9.13.2. Pressure Sensors

Manifold Absolute Pressure (MAP) Sensors

Figure 40. MAP Pressure Sensor Calibration Configuration

Fuel Rail Pressure Sensor

Figure 41. Fuel Rail Pressure Sensor Configuration

132

Procedure for MAP Sensors

1. An Omegadyne pressure transducer of known specifications was screwed into a pressure

vessel.

2. The Omegadyne pressure transducer was connected to a National Instruments data

acquisition module and the pressure from the pressure transducer was read from a

computer screen through LabView.

3. A ball valve was screwed into the pressure vessel to safely relieve pressure at the end of

each of the tests.

4. A compressed air cylinder was connected to the pressure vessel.

5. The Manifold Absolute Pressure (MAP) sensor was pushed into a tight gasket fitting in

the top of the pressure vessel and duct taped in place.

6. The wire leads from the MAP sensor were connected to a power supply and a multimeter.

Pin 1 on the MAP sensor was connected to the voltage terminal of the multimeter, pin 2

was connected to the positive terminal on the power supply and pin 4 was connected to

the negative terminal on the power supply and the common terminal of the multimeter.

7. The power supply was turned on and set to 5 volts.

8. The multimeter was turned on and set to DC voltage reading.

9. The valve on the compressed air cylinder was turned until the voltage reading on the

multimeter read each of the desired set points.

10. The pressure of the known pressure transducer, which was displayed by LabView, was

recorded.

11. Steps 9 and 10 were repeated for all of the desired voltage set points.

12. After 5 volts was achieved by the multimeter, the compressed air cylinder was closed and

the compressed air in the pressure vessel was relieved by slowly opening the ball valve

on the pressure vessel.

13. Steps 5 to 12 were repeated for each of the MAP sensors.

133

Procedure for Fuel Rail Pressure Sensor

1. An Omegadyne pressure transducer of known specifications was screwed into a pressure

vessel.

2. The pressure transducer was connected to a National Instruments data acquisition module

and the pressure from the pressure transducer was read from a computer screen through

LabView.

3. A ball valve was screwed into the pressure vessel to safely relieve pressure at the end of

each of the tests.

4. A compressed air cylinder was connected to the pressure vessel.

5. The fuel rail pressure sensor was screwed into the pressure vessel.

6. The wire leads from the fuel rail pressure sensor were connected to a power supply and a

multimeter. Pin 1 on the fuel rail pressure sensor was connected to the ground terminal of

the power supply and the common terminal of the multimeter, pin 2 was connected to the

positive terminal on the power supply and pin 3 was connected to the voltage terminal of

the multimeter.

7. The power supply was turned on and set to 5 volts.

8. The multimeter was turned on and set to DC voltage reading.

9. The valve on the compressed air cylinder was turned until the voltage reading on the

multimeter read each of the desired set points.

10. The pressure of the known pressure transducer, which was displayed by LabView, was

recorded.

11. Steps 9 and 10 were repeated for all of the desired voltage set points.

12. After 5 volts was achieved by the multimeter, the compressed air cylinder was closed and

the compressed air in the pressure vessel was relieved by slowly opening the ball valve

on the pressure vessel.

134

Results for MAP Sensors


, 2014

Pressure (psia) Voltage (V)

14.7 2.01

16.5 2.25

18.5 2.50

19.9 2.75

22.0 3.00

23.8 3.25

25.5 3.50

27.3 3.75

29.3 4.00

31.0 4.25

32.7 4.50

34.7 4.75

36.2 5.00

Table 10. Pressure Converted to kPa and MAP Sensor #1 Voltage Output

Pressure (kPa) Voltage (V)

101 2.01

114 2.25

128 2.50

137 2.75

152 3.00

164 3.25

176 3.50

188 3.75

202 4.00

214 4.25

225 4.50

239 4.75

250 5.00

135

Figure 42. Graph of Pressure vs. MAP Sensor #1 Voltage Output


, 2014


14.7 1.99

16.6 2.25

18.1 2.50

19.7 2.75

21.7 3.00

23.9 3.25

25.6 3.50

27.7 3.75

29.2 4.00

31.1 4.25

33.0 4.50

34.8 4.75

36.6 5.00

y = 49.80x + 1.80

0

50

100

150

200

250

300

0.00 1.00 2.00 3.00 4.00 5.00 6.00

Pre

ssu

re (

kPa)

MAP Sensor #1 Voltage Output (V)

Pressure (kPa) vs. MAP Sensor #1 Voltage Output (V)

136



101 1.99

114 2.25

125 2.50

136 2.75

150 3.00

165 3.25

177 3.50

191 3.75

201 4.00

214 4.25

228 4.50

240 4.75

252 5.00


y = 50.70x - 0.94

0

50

100

150

200

250

300

0.00 1.00 2.00 3.00 4.00 5.00 6.00

Pre

ssu

re (

kPa)



137


, 2014


14.7 1.99

16.6 2.25

18.4 2.50

20.0 2.75

22.1 3.00

23.6 3.25

25.7 3.50

27.6 3.75

29.4 4.00

31.1 4.25

33.0 4.50

34.7 4.75

36.7 5.00



101 1.99

114 2.25

127 2.50

138 2.75

152 3.00

163 3.25

177 3.50

190 3.75

203 4.00

214 4.25

228 4.50

239 4.75

253 5.00

138



, 2014


14.7 1.13

16.0 1.25

19.1 1.50

21.7 1.75

25.0 2.00

27.5 2.25

30.5 2.50

33.6 2.75

37.0 3.00

39.4 3.25

42.2 3.50

45.5 3.75

48.6 4.00

51.5 4.25

54.4 4.50

57.4 4.75

60.4 5.00

y = 50.36x + 0.70

0

50

100

150

200

250

300

0.00 1.00 2.00 3.00 4.00 5.00 6.00

Pre

ssu

re (

kPa)



139



101 1.13

110 1.25

132 1.50

150 1.75

172 2.00

190 2.25

210 2.50

232 2.75

255 3.00

272 3.25

291 3.50

314 3.75

335 4.00

355 4.25

375 4.50

396 4.75

416 5.00


y = 81.53x + 8.06

0

50

100

150

200

250

300

350

400

450

0.00 1.00 2.00 3.00 4.00 5.00 6.00

Pre

ssu

re (

kPa)



140

Fuel Rail Pressure Sensor

Table 17. Fuel Rail Pressure Sensor Voltage Output and Pressure from Test on August 13th

, 2014

Voltage (V) Pressure (psi)

0.53 14.7

0.75 22.4

1.00 31.0

1.25 39.6

1.50 48.3

1.75 57.0

2.00 65.5

2.25 74.2

2.50 83.3

2.75 91.5

3.00 101.0

3.25 109.3

3.50 117.8

3.75 126.5

4.00 135.1

4.25 143.8

4.50 152.5

4.75 162.0

5.00 170.7

141

Table 18. Fuel Rail Pressure Sensor Voltage Output and Pressure Converted to kPa

Voltage (V) Pressure (kPa)

0.53 101

0.75 154

1.00 214

1.25 273

1.50 333

1.75 393

2.00 452

2.25 512

2.50 574

2.75 631

3.00 696

3.25 754

3.50 812

3.75 872

4.00 931

4.25 991

4.50 1051

4.75 1117

5.00 1177

142

Figure 46. Graph of Pressure vs. Fuel Rail Pressure Sensor Voltage Output

Conclusion

All four of the MAP sensors displayed extremely linear relationships between pressure

and output voltage. Although some error can be expected from the known Omegadyne pressure

transducer, the error as a result of transient pressure changes inside the vessel and oscillation of

the displayed pressure were likely averaged out to zero with the linearization of the data. There

can be a high degree of confidence that the pressure read by the MAP sensors is very accurate.

The calibration could have been improved by using a more accurate known pressure transducer,

but given the precision of the MAP sensors, this gain would be marginal. Additionally, a

pressure vessel intended solely for the calibration of these MAP sensors would have reduced the

amount of air leaking from the apparatus and enabled both of the pressure transducers to be

exposed to the same pressure for longer. Again, this modification would have improved the lab

results very marginally. Although the pressure vessel used was not intended for the calibration of

these MAP sensors, the MAP sensors fit very tightly in their opening and very little air leaked

y = 240.12x - 27.04

0

200

400

600

800

1000

1200

1400

0.00 1.00 2.00 3.00 4.00 5.00 6.00

Pre

ssu

re (

kP

a)

Fuel Rail Pressure Sensor Voltage Output (V)

Pressure (kPa) vs. Fuel Rail Pressure Sensor

Voltage Output (V)

143

out. Moreover, the pressure sensors react very quickly to changes in pressure, so the effect of

slowly diminishing pressures would not greatly alter the results.

The resulting graph shows an extremely linear relation between pressure and sensor

output voltage across the entire sensor’s measuring range. Although the calibration lab was very

successful in terms of data gathering, some improvements could be made to the procedure in the

future. The lab was temporary, so the setup was not optimized for ease of use. In the future, the

regulator on the compressed air cylinder should be closer to the computer screen displaying the

Omegadyne pressure transducer pressure and the ball valve on the pressure valve. These three

elements that required frequent checking and adjustment were located far apart for this lab test,

which caused the operator to run around during the test. A more accurate pressure transducer

could have been used to calibrate the unknown pressure sensor, but given the precision of the

fuel rail pressure sensor, the results would be very marginally improved.

9.13.3. Temperature Sensors

Configuration for Air Temperature Tests

Figure 47. Air Temperature Sensor Configuration

144

Configuration for Ice Bucket Tests

Figure 48. Ice Bucket Configuration

Configuration for Heated Engine Coolant Bath

Figure 49. Heated Engine Coolant Bath Configuration

145

Procedure for Air Temperature Tests

1. An RTD probe of known specifications was lowered into the top of the Precision

Scientific oven.

2. The RTD probe was connected to a National Instruments data acquisition module and the

temperature of the RTD probe was read from a computer screen through LabView.

3. The wires of the unknown temperature sensor were wrapped in an insulating fabric to

prevent them from being damaged.

4. The temperature sensor of unknown specifications was lowered into the top of the

Precision Scientific oven and positioned next to the RTD probe.

5. A multimeter was connected to the temperature sensor of unknown specifications and set

to measure resistance.

6. The oven was turned on and left to heat up.

7. When the RTD probe read the desired temperature set point, the oven was turned off.

8. When the RTD probe indicated that the temperature had dropped below the set point, the

oven was turned back on.

9. Steps 7 and 8 were repeated as often as required such that the displayed temperature of

the RTD probe never drifted more than 0.25°C away from the desired set point

temperature. This process was performed for five minutes to ensure that the unknown

sensor was at equilibrium.

10. The resistance of the temperature sensor of unknown specifications was recorded.

11. Steps 6 to 10 were repeated for all of the desired temperature set points.

12. Steps 4 to 11 were repeated for all of the temperature sensors of unknown specifications.

Procedure for Ice Bucket Tests

1. A bucket was filled with equal parts ice and water.

2. The ice and water mixture was left for 20 minutes to reach a stable temperature of

approximately 0.0°C.

3. An RTD probe of known specifications was lowered into the bucket and duct taped in

place.

146



5. The unknown temperature sensor was lowered into the bucket next to the RTD probe and

duct taped in place.

6. A multimeter was connected to the unknown temperature sensor and set to measure

resistance.

7. The displayed temperature of the RTD probe was checked to ensure that the mixture was

0.0°C ±0.1°C and the resistance of the unknown temperature sensor was recorded.

8. Steps 5 to 7 were repeated for each of the unknown temperature sensors.

Procedure for Heated Engine Coolant Bath Tests

1. The PolyTemp constant temperature bath was filled with pure Prestone engine coolant so

that the heating coil in the bath was completely submerged in liquid.

2. An RTD probe of known specifications was lowered into the bath and duct taped in

place.



4. The unknown temperature sensor was lowered into the bath next to the RTD probe and

duct taped in place.

5. A multimeter was connected to the unknown temperature sensor and set to measure

resistance.

6. The power knob on the PolyTemp constant temperature bath was turned until the desired

temperature set point was reached, as indicated by the RTD probe.

7. The bath was left at each of the temperature set points for five minutes for the unknown

sensor to reach equilibrium and then its resistance was recorded.

8. Steps 6 and 7 were repeated for each of the desired temperature set points.

9. Steps 4 to 8 were repeated for each of the unknown temperature sensors.

147

Air Temperature Results


, 2014

Temperature (°C) MAP Sensor #1 Resistance (Ω)

30 1591

40 1157

50 861

60 664

70 491

80 388

90 299

100 226

110 173

120 134

130 102


, 2014


30 1808

40 1229

50 864

60 600

70 422

80 306

90 228

100 170

110 125

120 91

130 72

148

Table 21. Temperature and Resistance of MAP sensor #2 from Test on August 8th

, 2014


30 1750

40 1240

50 889

60 627

70 466

80 350

90 267

100 203

110 155

120 119

130 92


, 2014


30 1791

40 1206

50 819

60 588

70 425

80 308

90 234

100 176

110 124

120 94

130 72

149


, 2014


40 1300

50 944

60 735

70 540

80 414

90 325

100 219

110 164

120 124

130 98


, 2014


30 1811

40 1286

50 932

60 656

70 449

80 327

90 241

100 180

110 139

120 99

130 75

150


, 2014


30 1797

40 1332

50 970

60 667

70 451

80 329

90 240

100 178

110 133

120 99

130 76

Table 26. Temperature and Resistance of MAF Sensor from Test on August 18th

, 2014


30 25360

40 18960

50 13240

60 8710

70 5820

80 4080

90 3044

100 2196

110 1560

120 1133

130 870

151


, 2014


30 24660

40 17050

50 11450

60 7410

70 4860

80 3420

90 2477

100 1790

110 1330

120 982

130 755

Ice Bucket Results

Table 28. Resistance of Various Temperature Sensors at Zero Degrees from Test on August 19th

,

2014

Sensor Resistance (Ω)

Fuel Rail Temperature Sensor 84000

MAP Sensor #1 5540

MAP Sensor #2 5560

MAP Sensor #3 5420

MAP Sensor #4 5600

152

Heated Engine Coolant Bath Results


,

2014

Temperature (°C) Resistance (Ω)

45.5 4870

60.0 2790

70.0 1930

80.0 1380

90.0 1000

100.0 737

110.0 554

120.0 420

130.0 320

140.0 248

150.0 195


,

2014


0.0 37700

30.0 9230

40.0 6060

50.0 4090

60.0 2780

70.0 1950

80.0 1387

90.0 1002

100.0 737

110.0 549

120.0 416

130.0 319

140.0 248

150.0 195

153


,

2014


45.5 359.1

60.0 222.3

70.0 161.9

80.0 120.6

90.0 90.6

100.0 69.3

110.0 50.3

120.0 39.0

130.0 30.7

140.0 24.5

150.0 19.8


,

2014


0.0 2060

30.0 613

40.0 420

50.0 299

60.0 213

70.0 156

80.0 114

90.0 85.2

100.0 65.0

110.0 49.9

120.0 39.0

130.0 30.6

140.0 24.5

150.0 19.6

154


20th

, 2014


30.0 24660

40.0 16020

50.0 10740

60.0 7460

70.0 5270

80.0 3790

90.0 2778

100.0 2059


21st, 2014


40.0 16280

50.0 10820

60.0 7510

70.0 5290

80.0 3800

90.0 2786

100.0 2065

110.0 1554

120.0 1186

130.0 916

140.0 714

150.0 564

155

Combined Graphs

Figure 50. Graph of Temperature vs. Resistance of MAP Sensor #1


0.0

20.0

40.0

60.0

80.0

100.0

120.0

140.0

0 1000 2000 3000 4000 5000 6000

Tem

pe

ratu

re (

°C)

Resistance of MAP Sensor #1 (Ω)

Temperature (°C) vs. Resistance of MAP Sensor #1 (Ω)

0.0

20.0

40.0

60.0

80.0

100.0

120.0

140.0

0 1000 2000 3000 4000 5000 6000

Tem

pe

ratu

re (

°C)



156



0.0

20.0

40.0

60.0

80.0

100.0

120.0

140.0

0 1000 2000 3000 4000 5000 6000

Tem

pe

ratu

re (

°C)



0.0

20.0

40.0

60.0

80.0

100.0

120.0

140.0

0 1000 2000 3000 4000 5000 6000

Tem

pe

ratu

re (

°C)



157

Figure 54. Graph of Temperature vs. Resistance of Fuel Rail Temperature Sensor

Figure 55. Graph of Temperature vs. Resistance of MAF Sensor

0

20

40

60

80

100

120

140

160

0 10000 20000 30000 40000 50000 60000 70000 80000 90000

Tem

per

atu

re (°C

)

Resistance of Fuel Rail Temperature Sensor (Ω)

Temperature (°C) vs. Resistance of Fuel

Rail Temperature Sensor (Ω)

0

20

40

60

80

100

120

140

0 5000 10000 15000 20000 25000 30000

Tem

pe

ratu

re (

°C)

Resistance of MAF Sensor (Ω)

Temperature (°C) vs. Resistance of MAF Sensor (Ω)

158

Figure 56. Graph of Temperature vs. Resistance of Oil Temperature Sensor

Figure 57. Graph of Temperature vs. Resistance of Coolant Temperature Sensor

0.0

20.0

40.0

60.0

80.0

100.0

120.0

140.0

160.0

0 500 1000 1500 2000 2500

Tem

pe

ratu

re (

°C)

Resistance of Oil Temperature Sensor (Ω)

Temperature (°C) vs. Resistance of Oil Temperature Sensor (Ω)

0.0

20.0

40.0

60.0

80.0

100.0

120.0

140.0

160.0

0 5000 10000 15000 20000 25000 30000 35000 40000

Tem

pe

ratu

re (

°C)

Resistance of Coolant Temperature Sensor (Ω)

Temperature (°C) vs. Resistance of Coolant Temperature Sensor (Ω)

159

Conclusion

In comparison to the other calibration labs, several changes to the temperature calibration

lab could yield greatly improved results. The main challenge that was faced in this lab was

keeping a constant temperature for a long enough time that all of the temperature sensors of

different thermal masses reached equilibrium. The precision scientific oven that was used could

not hold at the low temperatures that were required for the calibration. This meant that the

operator of the lab needed to turn on and off the oven to keep a relatively constant temperature.

Although this process was effective for keeping the RTD sensor close to the desired set point, the

other sensors were likely oscillating above and below the set point with a large margin of error.

The other source of error for this lab that was probably significant was the distance between the

RTD sensor and the temperature sensors that were being calibrated. The height of the oven and

the length of the RTD probe wires meant that the RTD probe was quite far above the other

temperature sensors. Having the other sensors closer to the element of the oven, which was on

the bottom, meant that the temperature sensors being calibrated were likely hotter than the RTD

probe. This source of error was likely the largest in the lab.

Most of the sensors were meant for measuring the temperature of a gas or the temperature

of a liquid.

160

Filter Elements 9.14.



, 2016 on the Right



, 2016 on the Right

161



, 2016 on the Right


of 0.5 on March 3rd

, 2016 on the Right

162


of 0.6 on March 3rd

, 2016 on the Right


of 0.4 on March 3rd

, 2016 on the Right

163


of 0.5 on March 3rd

, 2016 on the Right



, 2016 on the Right

164



, 2016 on the Right



, 2016 on the Right

165



, 2016 on the Right



, 2016 on the Right

166

Figure 70. Unused Filter on the Left and 30 Minute Oven Test on the Right

Figure 71. Unused Filter on the Left and 30 Minute Isokinetic Test on the Right

167

Minimum Dilution Ratio Calculations 9.15.

Figure 72. Diagram of Dilution Streams

Humidity Ratio:

𝜔 =𝑚𝑣𝑎𝑝𝑜𝑢𝑟

𝑚𝑑𝑟𝑦 𝑎𝑖𝑟

Dilution Ratio:

𝐷𝑅 =�̇�𝑑𝑎

�̇�𝑒𝑥

At an equivalence ratio of 0.6 it can be shown that:

𝜒𝐻2𝑂 = 0.2233

𝜒𝑁2 = 0.7022

168

𝜒𝑂2 = 0.0744

𝜔𝑒𝑥 =𝜒𝐻2𝑂𝑀𝑊𝐻2𝑂

𝜒𝑁2𝑀𝑊𝑁2 + 𝜒𝑂2𝑀𝑊𝑂2

=(0.2233) (18.01528


)

(0.7022) (28.0134𝑘𝑔𝑘𝑚𝑜𝑙

) + (0.0744) (15.9994𝑘𝑔𝑘𝑚𝑜𝑙

)

𝜔𝑒𝑥 = 0.1928 𝑘𝑔(𝑣𝑎𝑝𝑜𝑢𝑟)

𝑘𝑔(𝑑𝑟𝑦 "𝑎𝑖𝑟")

𝜔𝑑𝑚 =𝜔𝑒𝑥�̇�𝑒𝑥 + 𝜔𝑑𝑎�̇�𝑑𝑎

�̇�𝑒𝑥 + �̇�𝑑𝑎=𝜔𝑒𝑥�̇�𝑒𝑥 + 𝜔𝑑𝑎𝐷𝑅�̇�𝑒𝑥

�̇�𝑒𝑥 + 𝐷𝑅�̇�𝑒𝑥=�̇�𝑒𝑥(𝜔𝑒𝑥 + 𝜔𝑑𝑎𝐷𝑅)

�̇�𝑒𝑥(1 + 𝐷𝑅)=𝜔𝑒𝑥 + 𝜔𝑑𝑎𝐷𝑅

1 + 𝐷𝑅

𝜔𝑑𝑚 + 𝜔𝑑𝑚𝐷𝑅 = 𝜔𝑒𝑥 +𝜔𝑑𝑎𝐷𝑅 → 𝜔𝑑𝑚𝐷𝑅 − 𝜔𝑑𝑎𝐷𝑅 = 𝜔𝑒𝑥 − 𝜔𝑑𝑚

𝐷𝑅(𝜔𝑑𝑚 − 𝜔𝑑𝑎) = 𝜔𝑒𝑥 − 𝜔𝑑𝑚 → 𝐷𝑅 =𝜔𝑒𝑥 − 𝜔𝑑𝑚𝜔𝑑𝑚 − 𝜔𝑑𝑎

If we use dilution air at 25°C, 10% relative humidity, then from Figure A-9, Pg. 994 of

(Moran, Shapiro, Boettner, & Bailey, 2011):

𝜔𝑑𝑎 = 0.002 𝑘𝑔(𝑣𝑎𝑝𝑜𝑢𝑟)


If we assume that the temperature of the diluted mixture is 25°C, which is reasonable

considering all of the fittings between the exhaust and the EEPS, then:

169

Table 35. Dilution Ratio Calculation for an Equivalence Ratio of 0.6

Relative Humidity (%)

Humidity Ratio (kg of

vapour/kg of “air”) [From

Figure A-9 Pg. 994 of (Moran,

Shapiro, Boettner, & Bailey,

2011)]

Dilution Ratio

100 0.02 9.6

80 0.016 12.6

60 0.012 18.1

40 0.008 30.8

20 0.004 94.4

Sample calculation for a relative humidity of 100%:

𝐷𝑅 =𝜔𝑒𝑥 − 𝜔𝑑𝑚𝜔𝑑𝑚 − 𝜔𝑑𝑎

=0.1928 − 0.02

0.02 − 0.002= 9.6

Table 36. Dilution Ratio Calculation for an Equivalence Ratio of 0.4




Figure A-9 Pg. 994 (Moran,


2011)]

Dilution Ratio

100 0.02 5.8

80 0.016 7.8

60 0.012 11.3

40 0.008 19.5

20 0.004 60.5



=(0.1547) (18.01528


)

(0.7294) (28.0134𝑘𝑔𝑘𝑚𝑜𝑙

) + (0.1160) (15.9994𝑘𝑔𝑘𝑚𝑜𝑙

)



170



=0.1250 − 0.02

0.02 − 0.002= 5.8

Table 37. Dilution Ratio for an Equivalence Ratio of 0.5




Figure A-9 Pg. 994 (Moran,


2011)]

Dilution Ratio

100 0.02 7.7

80 0.016 10.2

60 0.012 14.7

40 0.008 25.1

20 0.004 77.3



=(0.1897) (18.01528


)

(0.7155) (28.0134𝑘𝑔𝑘𝑚𝑜𝑙

) + (0.0948) (15.9994𝑘𝑔𝑘𝑚𝑜𝑙

)





=0.1585 − 0.02

0.02 − 0.002= 7.7

171

Figure 73. Psychrometric Chart (Moran, Shapiro, Boettner, & Bailey, 2011)

172

PM Correction to Raw Exhaust Gas Basis 9.16.

Correcting the number concentration from the EEPS to pure exhaust basis:

𝑃𝑀𝐸𝐸𝑃𝑆 =𝑃𝑀𝑒𝑥ℎ𝑎𝑢𝑠𝑡 + 𝐷𝑅 × 𝑃𝑀𝑑𝑖𝑙𝑢𝑡𝑖𝑜𝑛 𝑎𝑖𝑟

𝐷𝑅 + 1

𝑃𝑀𝑒𝑥ℎ𝑎𝑢𝑠𝑡 = (𝐷𝑅 + 1)𝑃𝑀𝐸𝐸𝑃𝑆 − 𝐷𝑅 × 𝑃𝑀𝑑𝑖𝑙𝑢𝑡𝑖𝑜𝑛 𝑎𝑖𝑟

The dilution ratio for all of the tests is 100.295 as described in section 3.5.2.

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