Gaseous and Particulate Matter Emissions of a Supercharged ... · that the current EPA regulations...
Transcript of Gaseous and Particulate Matter Emissions of a Supercharged ... · that the current EPA regulations...
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
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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.
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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.
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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
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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
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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.
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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
Table 7. Time between Actuation of Micro Switches and Voltage Output of MAF Sensor from
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
Table 11. Pressure and MAP Sensor #3 Voltage Output from Test on August 12th
, 2014 ........ 135
Table 12. Pressure Converted to kPa and MAP Sensor #3 Voltage Output ............................... 136
Table 13. Pressure and MAP Sensor #4 Voltage Output from Test on August 12th
, 2014 ........ 137
Table 14. Pressure Converted to kPa and MAP Sensor #4 Voltage Output ............................... 137
Table 15. Pressure and MAP Sensor #2 Voltage Output from Test on August 12th
, 2014 ........ 138
Table 16. Pressure Converted to kPa and MAP Sensor #2 Voltage Output ............................... 139
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
Table 20. Temperature and Resistance of MAP Sensor #1 from Test on August 18th
, 2014 ..... 147
Table 21. Temperature and Resistance of MAP sensor #2 from Test on August 8th
, 2014 ........ 148
Table 22. Temperature and Resistance of MAP Sensor #2 from Test on August 18th
, 2014 ..... 148
Table 23. Temperature and Resistance of MAP Sensor #3 from Test on August 8th
, 2014 ....... 149
Table 24. Temperature and Resistance of MAP Sensor #3 from Test on August 18th
, 2014 ..... 149
Table 25. Temperature and Resistance of MAP Sensor #4 from Test on August 18th
, 2014 ..... 150
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Table 26. Temperature and Resistance of MAF Sensor from Test on August 18th
, 2014 .......... 150
Table 27. Temperature and Resistance of MAP Sensor #4 from Test on August 18th
, 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
Table 30. Temperature and Resistance of Coolant Temperature Sensor from Test on August 18th
,
2014............................................................................................................................................. 152
Table 31. Temperature and Resistance of Oil Temperature Sensor from Test on August 15th
,
2014............................................................................................................................................. 153
Table 32. Temperature and Resistance of Oil Temperature Sensor from Test on August 18th
,
2014............................................................................................................................................. 153
Table 33. Temperature and Resistance of Fuel Rail Temperature Sensor from Test on August
20th
, 2014 .................................................................................................................................... 154
Table 34. Temperature and Resistance of Fuel Rail Temperature Sensor from Test on August
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
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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 17. NOx and Fuel Conversion Efficiency vs. Ignition Timing for φ = 0.5 ....................... 57
Figure 18. NOx and Fuel Conversion Efficiency vs. Ignition Timing for φ = 0.6 ....................... 58
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
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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 43. Graph of Pressure vs. MAP Sensor #3 Voltage Output ............................................ 136
Figure 44. Graph of Pressure vs. MAP Sensor #4 Voltage Output ............................................ 138
Figure 45. Graph of Pressure vs. MAP Sensor #2 Voltage Output ............................................ 139
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
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Figure 50. Graph of Temperature vs. Resistance of MAP Sensor #1 ......................................... 155
Figure 51. Graph of Temperature vs. Resistance of MAP Sensor #2 ......................................... 155
Figure 52. Graph of Temperature vs. Resistance of MAP Sensor #3 ......................................... 156
Figure 53. Graph of Temperature vs. Resistance of MAP Sensor #4 ......................................... 156
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
Figure 59. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.5 on February 26th
, 2016 on the Right ................................................................................. 160
Figure 60. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.6 on February 26th
, 2016 on the Right ................................................................................. 161
Figure 61. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.5 on March 3rd
, 2016 on the Right ....................................................................................... 161
Figure 62. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.6 on March 3rd
, 2016 on the Right ....................................................................................... 162
Figure 63. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.4 on March 3rd
, 2016 on the Right ....................................................................................... 162
Figure 64. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
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
Figure 66. 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 ............................................................................. 164
Figure 67. Unused Filter on the Left and 30 Minute Naturally Aspirated Test at an Equivalence
Ratio of 0.6 on March 8th
, 2016 on the Right ............................................................................. 164
Figure 68. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.6 on March 10th
, 2016 on the Right ..................................................................................... 165
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Figure 69. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.4 on March 10th
, 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
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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
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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
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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
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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 12, 2016 0.5 Supercharged Spark Timing Test
February 12, 2016 0.4 Supercharged Spark Timing Test
February 26, 2016 0.4 Supercharged Fixed Spark Timing
February 26, 2016 0.5 Supercharged Fixed Spark Timing
February 26, 2016 0.6 Supercharged Fixed Spark Timing
March 3, 2016 0.5 Supercharged Fixed Spark Timing
March 3, 2016 0.6 Supercharged Fixed Spark Timing
March 3, 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
March 5, 2016 0.6 Naturally Aspirated Fixed Spark Timing
March 8, 2016 0.6 Naturally Aspirated Fixed Spark Timing
March 10, 2016 0.6 Supercharged Fixed Spark Timing
March 10, 2016 0.4 Supercharged 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. NOx and Fuel Conversion Efficiency vs. Ignition Timing for φ = 0.5
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
)
Ignition Timing (°BTDC)
Emissions Bench NOx (ppm) and Fuel Conversion Efficiency (%) vs. Ignition Timing (°BTDC) for Spark Timing Tests and Full Tests at φ = 0.5
N0x Spark Timing Tests SC
N0x Full Tests SC
η_fc Spark Timing Tests SC
η_fc Full Tests SC
58
Figure 18. NOx and Fuel Conversion Efficiency vs. Ignition Timing for φ = 0.6
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
)
Ignition Timing (°BTDC)
Emissions Bench NOx (ppm) and Fuel Conversion Efficiency (%) vs. Ignition Timing (°BTDC) for Spark Timing Tests and Full Tests at φ = 0.6
N0x Spark Timing Tests SC
N0x Full Tests SC
η_fc Spark Timing 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
)
Emissions Bench O₂ Calculated Equivalence Ratio
NOx produced per km (mg·NOx/km) vs. Emissions Bench O₂ Calculated Equivalence Ratio
Supercharged Tests
Naturally Aspirated 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 (
%)
Emissions Bench O₂ Calculated Equivalence Ratio
Fuel Conversion Efficiency (%) vs. Emissions Bench O₂ Calculated Equivalence Ratio
Supercharged Tests
Naturally Aspirated 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)
Emissions Bench O₂ Calculated Equivalence Ratio
FTIR Calculated Lubricating Oil Consumption Rate (g/hr) vs. Emissions Bench O₂ Calculated
Equivalence Ratio
Supercharged Tests
Naturally Aspirated 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
Naturally Aspirated 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) =
�̇�𝑎𝑖𝑟
�̇�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
�̇�𝑎𝑖𝑟
= {
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
𝑀𝑊𝑒𝑥ℎ𝑎𝑢𝑠𝑡
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
𝑀𝑊𝑒𝑥ℎ𝑎𝑢𝑠𝑡
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 +
3.773 − 4.773𝜒𝐶𝑂2,𝑏𝑎𝑐𝑘𝑔𝑟𝑜𝑢𝑛𝑑2𝜑
}
1 − 𝜒𝐶𝑂2,𝑙𝑖𝑐𝑜𝑟
}
(0.5𝜑+ 0.5 +
3.773 − 4.773𝜒𝐶𝑂2,𝑏𝑎𝑐𝑘𝑔𝑟𝑜𝑢𝑛𝑑2𝜑
+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
𝑀𝑊𝑒𝑥ℎ𝑎𝑢𝑠𝑡
𝜒𝐶𝑂,𝑒𝑥ℎ𝑎𝑢𝑠𝑡 = (𝑀𝑊𝑒𝑥ℎ𝑎𝑢𝑠𝑡
𝑀𝑊𝑖𝑛𝑡𝑎𝑘𝑒) 𝜒𝐶𝑂,𝑖𝑛𝑡𝑎𝑘𝑒
𝜒𝐶𝑂,𝑒𝑥ℎ𝑎𝑢𝑠𝑡 = (26.07712897
𝑘𝑔𝑘𝑚𝑜𝑙
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();
}
else if(user_input =='3')
{
OneStepBackward();
}
else if(user_input =='4')
{
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"
for(x= 0; x<10; 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 OneStepBackward()
{
Serial.println("Moving One Backward");
digitalWrite(2, LOW); //Pull direction pin low to move "backward"
for(x= 0; x<1; x++) //Loop the forward stepping enough times for motion to be visible
{
digitalWrite(3,HIGH); //Trigger one step backward
delay(1);
digitalWrite(3,LOW); //Pull step pin low so it can be triggered again
delay(1);
}
Serial.println("Enter New Command");
Serial.println();
}
115
void TenStepsBackward()
{
Serial.println("Moving Ten Backward");
digitalWrite(2, LOW); //Pull direction pin low to move "backward"
for(x= 0; x<10; x++) //Loop the forward stepping enough times for motion to be visible
{
digitalWrite(3,HIGH); //Trigger one step backward
delay(1);
digitalWrite(3,LOW); //Pull step pin low so it can be triggered again
delay(1);
}
Serial.println("Enter New Command");
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
Table 6. Time between Actuation of Micro Switches and Voltage Output of MAF Sensor from
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
Table 7. Time between Actuation of Micro Switches and Voltage Output of MAF Sensor from
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
Table 9. Pressure and MAP Sensor #1 Voltage Output from Test on August 12th
, 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
Table 11. Pressure and MAP Sensor #3 Voltage Output from Test on August 12th
, 2014
Pressure (psia) Voltage (V)
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
Table 12. Pressure Converted to kPa and MAP Sensor #3 Voltage Output
Pressure (kPa) Voltage (V)
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
Figure 43. Graph of Pressure vs. MAP Sensor #3 Voltage Output
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)
MAP Sensor #2 Voltage Output (V)
Pressure (kPa) vs. MAP Sensor #3 Voltage Output (V)
137
Table 13. Pressure and MAP Sensor #4 Voltage Output from Test on August 12th
, 2014
Pressure (psia) Voltage (V)
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
Table 14. Pressure Converted to kPa and MAP Sensor #4 Voltage Output
Pressure (kPa) Voltage (V)
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
Figure 44. Graph of Pressure vs. MAP Sensor #4 Voltage Output
Table 15. Pressure and MAP Sensor #2 Voltage Output from Test on August 12th
, 2014
Pressure (psia) Voltage (V)
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)
MAP Sensor #3 Voltage Output (V)
Pressure (kPa) vs. MAP Sensor #4 Voltage Output (V)
139
Table 16. Pressure Converted to kPa and MAP Sensor #2 Voltage Output
Pressure (kPa) Voltage (V)
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
Figure 45. Graph of Pressure vs. MAP Sensor #2 Voltage Output
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)
MAP Sensor #4 Voltage Output (V)
Pressure (kPa) vs. MAP Sensor #2 Voltage Output (V)
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
4. 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.
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.
3. 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.
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
Table 19. Temperature and Resistance of MAP Sensor #1 from Test on August 8th
, 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
Table 20. Temperature and Resistance of MAP Sensor #1 from Test on August 18th
, 2014
Temperature (°C) MAP Sensor #1 Resistance (Ω)
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
Temperature (°C) MAP Sensor #2 Resistance (Ω)
30 1750
40 1240
50 889
60 627
70 466
80 350
90 267
100 203
110 155
120 119
130 92
Table 22. Temperature and Resistance of MAP Sensor #2 from Test on August 18th
, 2014
Temperature (°C) MAP Sensor #2 Resistance (Ω)
30 1791
40 1206
50 819
60 588
70 425
80 308
90 234
100 176
110 124
120 94
130 72
149
Table 23. Temperature and Resistance of MAP Sensor #3 from Test on August 8th
, 2014
Temperature (°C) MAP Sensor #3 Resistance (Ω)
40 1300
50 944
60 735
70 540
80 414
90 325
100 219
110 164
120 124
130 98
Table 24. Temperature and Resistance of MAP Sensor #3 from Test on August 18th
, 2014
Temperature (°C) MAP Sensor #3 Resistance (Ω)
30 1811
40 1286
50 932
60 656
70 449
80 327
90 241
100 180
110 139
120 99
130 75
150
Table 25. Temperature and Resistance of MAP Sensor #4 from Test on August 18th
, 2014
Temperature (°C) MAP Sensor #4 Resistance (Ω)
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
Temperature (°C) MAP Sensor #1 Resistance (Ω)
30 25360
40 18960
50 13240
60 8710
70 5820
80 4080
90 3044
100 2196
110 1560
120 1133
130 870
151
Table 27. Temperature and Resistance of MAP Sensor #4 from Test on August 18th
, 2014
Temperature (°C) MAP Sensor #4 Resistance (Ω)
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
Table 29. Temperature and Resistance of Coolant Temperature Sensor from Test on August 15th
,
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
Table 30. Temperature and Resistance of Coolant Temperature Sensor from Test on August 18th
,
2014
Temperature (°C) Resistance (Ω)
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
Table 31. Temperature and Resistance of Oil Temperature Sensor from Test on August 15th
,
2014
Temperature (°C) Resistance (Ω)
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
Table 32. Temperature and Resistance of Oil Temperature Sensor from Test on August 18th
,
2014
Temperature (°C) Resistance (Ω)
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
Table 33. Temperature and Resistance of Fuel Rail Temperature Sensor from Test on August
20th
, 2014
Temperature (°C) Resistance (Ω)
30.0 24660
40.0 16020
50.0 10740
60.0 7460
70.0 5270
80.0 3790
90.0 2778
100.0 2059
Table 34. Temperature and Resistance of Fuel Rail Temperature Sensor from Test on August
21st, 2014
Temperature (°C) Resistance (Ω)
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
Figure 51. Graph of Temperature vs. Resistance of MAP Sensor #2
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)
Resistance of MAP Sensor #2 (Ω)
Temperature (°C) vs. Resistance of MAP Sensor #2 (Ω)
156
Figure 52. Graph of Temperature vs. Resistance of MAP Sensor #3
Figure 53. Graph of Temperature vs. Resistance of MAP Sensor #4
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 #3 (Ω)
Temperature (°C) vs. Resistance of MAP Sensor #3 (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)
Resistance of MAP Sensor #4 (Ω)
Temperature (°C) vs. Resistance of MAP Sensor #4 (Ω)
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.
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
Figure 59. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.5 on February 26th
, 2016 on the Right
161
Figure 60. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.6 on February 26th
, 2016 on the Right
Figure 61. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.5 on March 3rd
, 2016 on the Right
162
Figure 62. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.6 on March 3rd
, 2016 on the Right
Figure 63. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.4 on March 3rd
, 2016 on the Right
163
Figure 64. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.5 on March 3rd
, 2016 on the Right
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
164
Figure 66. 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
Figure 67. Unused Filter on the Left and 30 Minute Naturally Aspirated Test at an Equivalence
Ratio of 0.6 on March 8th
, 2016 on the Right
165
Figure 68. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.6 on March 10th
, 2016 on the Right
Figure 69. Unused Filter on the Left and 20 Minute Supercharged Test at an Equivalence Ratio
of 0.4 on March 10th
, 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
𝑘𝑔𝑘𝑚𝑜𝑙
)
(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)
𝑘𝑔(𝑑𝑟𝑦 "𝑎𝑖𝑟")
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
Relative Humidity (%)
Humidity Ratio (kg of
vapour/kg of “air”) [From
Figure A-9 Pg. 994 (Moran,
Shapiro, Boettner, & Bailey,
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
𝜔𝑒𝑥 =𝜒𝐻2𝑂𝑀𝑊𝐻2𝑂
𝜒𝑁2𝑀𝑊𝑁2 + 𝜒𝑂2𝑀𝑊𝑂2
=(0.1547) (18.01528
𝑘𝑔𝑘𝑚𝑜𝑙
)
(0.7294) (28.0134𝑘𝑔𝑘𝑚𝑜𝑙
) + (0.1160) (15.9994𝑘𝑔𝑘𝑚𝑜𝑙
)
𝜔𝑒𝑥 = 0.1250 𝑘𝑔(𝑣𝑎𝑝𝑜𝑢𝑟)
𝑘𝑔(𝑑𝑟𝑦 "𝑎𝑖𝑟")
170
Sample calculation for a relative humidity of 100%:
𝐷𝑅 =𝜔𝑒𝑥 − 𝜔𝑑𝑚𝜔𝑑𝑚 − 𝜔𝑑𝑎
=0.1250 − 0.02
0.02 − 0.002= 5.8
Table 37. Dilution Ratio for an Equivalence Ratio of 0.5
Relative Humidity (%)
Humidity Ratio (kg of
vapour/kg of “air”) [From
Figure A-9 Pg. 994 (Moran,
Shapiro, Boettner, & Bailey,
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
𝜔𝑒𝑥 =𝜒𝐻2𝑂𝑀𝑊𝐻2𝑂
𝜒𝑁2𝑀𝑊𝑁2 + 𝜒𝑂2𝑀𝑊𝑂2
=(0.1897) (18.01528
𝑘𝑔𝑘𝑚𝑜𝑙
)
(0.7155) (28.0134𝑘𝑔𝑘𝑚𝑜𝑙
) + (0.0948) (15.9994𝑘𝑔𝑘𝑚𝑜𝑙
)
𝜔𝑒𝑥 = 0.1585 𝑘𝑔(𝑣𝑎𝑝𝑜𝑢𝑟)
𝑘𝑔(𝑑𝑟𝑦 "𝑎𝑖𝑟")
Sample calculation for a relative humidity of 100%:
𝐷𝑅 =𝜔𝑒𝑥 − 𝜔𝑑𝑚𝜔𝑑𝑚 − 𝜔𝑑𝑎
=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.