Reactivity Reduced Fuel Reactivity Controlled

0 downloads 0 Views 11MB Size Report
Endress-Hauser Promass A. (SN#3N547253) ...... 0.40. 0.42. 0.44. 0.46. 170. 180. 190. 200. 210. 220. 5. 7. 9. 11. 13. 15. 17. -16 -14 -12 -10. -8 ... th e rm a l in d ic a te d. Injection Timing (deg ATDC). 70. 40. 55. 55. IS. F. C g a s o lin e e q . n e.

i

Abstract An Experimental investigation of fuel reactivity controlled combustion was performed on a Caterpillar SCOTE 3401E heavy-duty engine. The experiments were designed to investigate the combustion process and fuel effects of the reactivity gradient. The investigation consisted of both in-cylinder spectra based optical diagnostics and indicated pressure measurements. Optical diagnostics were used to explore dual fuel operation with gasoline port fueling and diesel direct injection. Further experimentation without the optical diagnostics was conducted with two alternative fueling strategies. The first strategy used gasoline port fueling and direct-injection of gasoline blended with a cetane improver. The second strategy used port fueling of E-98 and direct injection of diesel fuel. The optical diagnostics were used to measure natural thermal emission spectra of combustion in the mid-infrared broad-band at two measurement locations. The engine was operated at a light load of 4.5 bar indicated mean effective pressure (IMEP). Qualitative species trends and reaction extents of measured fuel, water, and combustion intermediates were compared to multidimensional KIVA-CFD simulations, with good agreement seen at both locations. The diagnostics demonstrated that the combustion process is controlled by the reactivity gradient, where combustion initiates globally in areas of higher fuel reactivity, and progresses to areas of lower fuel reactivity. This extension of the premixed combustion regime was demonstrated to offer high thermal efficiency and while meeting EPA 2010 HD emissions mandates of NOx and PM in-cylinder.

ii Using the combustion process insight gained from the optical diagnostics, fuel effects on the reactivity gradient were studied with a cetane improver at loads of 6 and 9 bar IMEP. The cetane improver di-tert-butyl peroxide (DTBP) was added to gasoline at various addition levels, and direct injected. In these tests the same gasoline without cetane improver was port fuel injected. The results of this fueling strategy demonstrated that approximately 1.75% DTBP added to only the direct injected gasoline established a reactivity gradient similar to that provided by diesel. The emissions and performance of these tests were comparable to gasoline diesel fueling strategies previously demonstrated by Hanson [48]. The last series of tests explored the effect of increasing the size of the reactivity gradient through using ethanol was as the port fuel instead of gasoline. These tests were only conducted at a load of 9 bar IMEP. Similar to the optical tests, these tests also used diesel fuel as the direct injected fuel. It was found that operation with ethanol port fuel was significantly different than gasoline port fuel. The ethanol results demonstrated longer combustion duration, thus reducing EGR requirements. Also, it was also found that significantly more diesel (reactive fuel) quantity was needed with ethanol. These tests showed that fuel effects are critical to the combustion process, and that changes to the reactivity gradient alter the combustion event. The findings of these three of experiments were that combustion initiated in areas of high reactivity and progressed to areas of lower reactivity, and that the reactivity gradient and fuel effects are critical to the combustion event.

iii

Acknowledgements This research would not have been possible without the support and guidance of many individuals. Specifically I would like to thank Dr. Rolf Reitz for extending me the opportunity to be his student and work in his laboratory, and the Department of Energy for funding this research. Within the Engine Research Center there are several other individuals who have also been responsible for the success of this research. In particular the fellow graduate students Reed Hanson, Sage Kokjohn, Keith Rein, Mike Tess, Luke Staples, and Eric Weninger have all been instrumental in conducting and understanding of this research and general laboratory procedures. I would also like to thank Ralph Braun, Jacob Backhaus for laboratory instrumentation and fabrication and Caterpillar for fabrication of the optical cylinder head and hardware support. I would like to thank my parents for teaching me to peruse my goals and dreams. Most of all I would like to thank my wife Jessica for her guidance and support both in my research and life.

iv

Table of Contents Chapter 1: 1.1

Introduction .................................................................................17

Background ..........................................................................................17

1.1.1

Spark ignition engine fundamentals ..............................................18

1.1.2

Compression Ignition fundamentals ..............................................19

1.1.3

Advanced Combustion Strategies .................................................20

1.1.4

Research objective ........................................................................21

Chapter 2: 2.1

Literature Review.........................................................................22

Conventional Diesel (Compression Ignition) Combustion.....................22

2.1.1

Indirect injection systems ..............................................................22

2.1.2

Direct injection engines .................................................................23

2.1.3

Compression ignition emissions ....................................................26

2.1.4

Low Temperature Combustion (LTC) ............................................30

2.1.5

Homogeneous Charge Compression Ignition (HCCI)....................31

2.1.6

HCCI operation..............................................................................31

2.1.7

HCCI research...............................................................................32

2.2

Hydrocarbon Oxidation .........................................................................35

2.2.1

Fuel Decomposition Steps.............................................................35

2.3

Cetane Improvers .................................................................................38

2.4

Fuel reactivity stratification for extended duration fully premixed

combustion ......................................................................................................45 Chapter 3:

Experimental design ...................................................................48

v 3.1

Engine ..................................................................................................48

3.2

Injection Systems .................................................................................51

3.2.1

Low Pressure Injection Hardware and Plumbing ...........................51

3.2.2

High Pressure Injection Hardware and Plumbing ..........................54

3.2.3

Electronic Injection Control............................................................58

3.2.4

Electronic Common Rail Pressure Control ....................................60

3.3

Fuel Properties .....................................................................................62

3.4

Data Acquisition and Laboratory Hardware ..........................................63

3.5

Engine Out Emissions ..........................................................................65

3.6

Optics ...................................................................................................66

3.6.1

Optical access ...............................................................................66

3.6.2

Optical technique and hardware ....................................................68

Chapter 4: 4.1

Optical Investigation Results and Discussion ..........................73

Optical diagnostics ...............................................................................73

4.1.1

Experimental operating conditions ................................................73

4.1.2

Experimental and computational agreement baseline ...................75

4.1.3

Measured Species.........................................................................79

4.1.4

Broadband Spectra........................................................................80

4.1.5

Fuel Spectra ..................................................................................83

4.1.6

Intermediate species measurements.............................................84

4.1.7

Combustion product species measurements.................................86

4.1.8

Comparison Between Computational Simulation and

Measurements .............................................................................................87

vi 4.2

Spectra Temperature Compensation and Reaction Extent...................92

4.2.1

Aldehyde reaction extent comparison............................................96

4.2.2

Fuel reaction extent comparison ...................................................98

4.2.3

Water reaction extent comparison ...............................................100

4.2.4

Reaction extents and combustion trends.....................................102

Chapter 5:

Fuel Reactivity Controlled Combustion via Reactivity

enhancement...................................................................................................106 5.1

Reactivity improving chemicals...........................................................106

5.2

DTBP addition to Direct Injected Gasoline..........................................108

5.3

DTBP addition to achieve fuel reactivity stratification .........................112

5.3.1

Low Load DTBP Study ................................................................112

5.3.2

DTBP percent sweep with 0% EGR ............................................113

5.3.3

EGR Sweep.................................................................................119

5.3.4

9 (bar) operation ..........................................................................122

5.3.5

3.5% DTBP addition ....................................................................124

5.3.6

1.75% DTBP addition ..................................................................130

5.3.7

0.75% DTBP addition ..................................................................135

5.3.8

89% and 90% port fuel combustion comparisons........................139

5.3.9

Operation with DTBP and Increased EGR level ..........................141

Chapter 6:

Reactivity Reduced Fuel Reactivity Controlled Combustion 146

6.1

Reactivity reduction ............................................................................146

6.2

Ethanol Direct Injection Ignition Delay ................................................146

6.3

Fuel Reactivity Controlled Combustion with E-98 Ethanol and Diesel 153

vii 6.3.1 Chapter 7: 7.1

9 (bar) operation with E-98 port fuel ............................................153 Conclusions ...............................................................................166

Summary of Investigations .................................................................166

7.1.1

Experimental Conclusions ...........................................................166

7.1.2

Suggestions for Future Work.......................................................168

References……………………………………………………...……………………170 Appendix A: Part Drawings and Prints...……………………………………….177 Appendix B: Fuel Properties……………………………………………………...195 Appendix C: Engine Data……………………………………………………........200

viii

List of Figures Figure 1 Mixing controlled combustion jet development in crank angle degrees after start of injection (ASI) units, [5] ...................................................................26 Figure 2 Equivalence ratio temperature plot (Sun [13]).......................................28 Figure 3 Fully developed mixing controlled combustion jet [5]............................29 Figure 4 Contribution matrix of a reduced reaction set leading to a four step hydrocarbon oxidation system [29]. ....................................................................36 Figure 6 EPA study of cetane improvement to diesel fuel [37] using the correlations from [36] ..........................................................................................41 Figure 7 Rapid Compression Machine (RCM) ignition characteristics of PRF blends with 2% by volume of 2-EHN and DTBP. Note that the DTBP is more effective with a PRF representative of gasoline than 2-EHN, opposite to the behavior with diesel fuels [38] .............................................................................42 Figure 8 Chemical mechanism of DTBP as proposed by [42], note the scission of the peroxide bond occurs first.............................................................................43 Figure 9 Ignition delay and burn rate of various cetane improves vs. initial temperature. Note that cetane improvers have little to no effect on burn duration Tanka el al. [38] ..................................................................................................44 Figure 10 Combustion luminosity of high and low cetane diesel and low cetane diesel with 2-EHN cetane improver. Note that the cetane improver has no effect on luminosity trends, it only advances the time of initial luminosity [43]..............44 Figure 11 Diagram of the engine lab...................................................................49 Figure 12 low pressure fuel system adapted from Hanson [48] ..........................52

ix Figure 13 installed port injector in intake elbow ..................................................53 Figure 14 Common Rail Fuel Delivery System adapted from Hanson [48] .........54 Figure 15 Common rail adaptor insert assembly for use with HEUI 315 inserts .56 Figure 16 Bosch Common rail injector installed into HEUI 315B head with custom insert (purple) and clamps (orange and green)...................................................57 Figure 18 Common Rail Pressure Transducer Calibration..................................61 Figure 19 MotoTune VGS User Interface............................................................62 Figure 20 Analog signal schematic, as documented by Hardy [13] ....................64 Figure 21 Serial cable communication schematic, as documented by Hardy [13] ............................................................................................................................64 Figure 22 Gaseous emissions sampling system, adapted from Lichety [53].......66 Figure 26 Diagram of optical setup .....................................................................70 Figure 28 Experimental and computational simulation (Kokjohn [66]) of cylinder pressure and AHRR............................................................................................78 Figure 32 Location A (bowl) measured combustion products of water and carbon dioxide emission before (bottom) and after (top) combustion (units of emission are arbitrary). ......................................................................................................87 Figure 33 Select crank-angles of measured spectra and predicted species (units of emission are arbitrary). ...................................................................................91 Figure 34 Temperature power (n) dependence calculation where the numbers below each data point correspond to the crank-angle of each integrated fuel emission value (units of emission are arbitrary). .................................................93

x Figure 35 Location A (bowl) raw measured (bottom plot) and temperaturecorrected spectra using CFD simulation temperatures (top plot) (units of emission are arbitrary). ......................................................................................................94 Figure 36 Temperature-corrected fuel and aldehyde spectra at Location A (bowl). Only spectra from even crank angles are shown (the units of emission are arbitrary). ............................................................................................................95 Figure 37 Temperature-corrected fuel and aldehyde spectra at Location B (squish). Only spectra from even crank angles are shown, note the delay in combustion as opposed to location A (units of emission are arbitrary). ..............95 Figure 38 Comparison of reaction extent between experimentally measured aldehydes (solid lines) and KIVA CFD predicted formaldehyde (dashed lines). .97 Figure 39 Computational and experimental fuel reaction extent at both Locations A (bowl) and B (squish).......................................................................................99 Figure 40 Computational and experimental water reaction extent at both Locations A (bowl) and B (squish). ...................................................................101 Figure 41 CFD species mass-fraction predictions of fuel decomposition globally, and at Locations A (Kokjohn [66]). ....................................................................102 Figure 42 Experimental Cylinder pressure, AHRR, and experimental optical Location A (bowl) and B (squish) aldehyde, fuel, and water reaction extent measurements. .................................................................................................104 Figure 43 HCCI ignition delay and fueling requirement as a function of % DTBP addition to iso-octane [39]. ................................................................................107

xi Figure 44 Direct injection gasoline and DTBP. Ignition delay numbers indicate intake temperature. ...........................................................................................109 Figure 45 Combustion timing and efficiency of direct injection of gasoline at 3 (bar) IMEP operation both with and without 2% DTBP addition. .......................110 Figure 48 Pressure and AHRR traces for in-cylinder blended fuel reactivity stratification via DTBP addition to pump gasoline at 6 (bar) load......................117 Figure 49 Combustion behavior trends of various DTBP percentage additions to gasoline, for the tests in Table 15. ....................................................................119 Figure 50 EGR sweep with 3.5% DTBP at 6 (bar) net IMEP ............................120 Figure 52 Indicated pressure and AHRR trends of 3.5% DTBP addition to direct injected gasoline in fuel reactivity controlled combustion at 9 (bar) IMEP, compared to the highest efficiency case of Hanson [47, 48], (black). ...............125 Figure 53 Combustion trend of 3.5% DTBP addition to direct injected gasoline in fuel reactivity controlled combustion at 9 (bar) IMEP, as compared to cases of Hanson [47, 48], (black)....................................................................................125 Figure 55 Indicated pressure and AHRR trends of 1.75% DTBP addition to directinjected gasoline in fuel reactivity-controlled combustion at 9 (bar) IMEP, compared to the highest efficiency case of Hanson [47, 48], (black). ...............130 Figure 56 Comparison of 3.5% (solid colored lines) and 1.75% (dashed colored lines) DTBP with identical port-fueling percentages, and compared to the highest efficiency case of Hanson [47, 48], (black). ......................................................132

xii Figure 57 Combustion trend of 3.5% and 1.75% DTBP addition to direct injected gasoline in fuel reactivity-controlled combustion at 9 (bar) IMEP, compared to cases of Hanson [47, 48]. .................................................................................133 Figure 59 Indicated pressure and AHRR trends of 0.75% DTBP addition to directinjected gasoline in fuel reactivity-controlled combustion at 9 (bar) IMEP, compared to the highest efficiency case of Hanson [47, 48], (black). ...............136 Figure 60 Combustion trends of 3.5%, 1.75%, and 0.75% DTBP addition to direct-injected gasoline in fuel reactivity-controlled combustion at 9 (bar) IMEP, compared to cases of Hanson [47, 48]. ............................................................136 Figure 62 Identical port fueling percentages between gasoline with DTBP and diesel of [47, 48] (black), at 9 (bar) IMEP..........................................................139 Figure 63 Similar positions of main heat release between gasoline/gasoline with DTBP and gasoline/diesel [47, 48] (black), shown at 9 (bar) IMEP...................141 Figure 64 Cylinder pressure and AHRR of 3.5% DTBP addition to direct injected gasoline with 43% EGR (solid colored lines) and 50% EGR (dotted colored lines), and 9 (bar) IMEP compared to the gasoline/diesel operation of Hanson [47, 48] 43% EGR (black). .............................................................................................142 Figure 65 Combustion trends of 3.5% DTBP with 43 and 50% EGR and gasoline/diesel operation of Hanson [47, 48] with 43% EGR. ...........................143 Figure 67 Ignition delay trends of all gasoline-based fuels tested with the operating conditions shown Table 13, number indicates intake temperature....149 Figure 68 Combustion trends of all gasoline-based fuels tested with the operating conditions shown in Table 13, number indicates intake temperature................149

xiii Figure 70 Indicated cylinder pressure and AHRR for E-98 port fueling, with constant main heat release peak timing and 9 (bar) IMEP operation................154 Figure 71 Indicated cylinder pressure and AHRR of DTBP/gasoline, diesel/gasoline, and diesel/E-98 fuel reactivity-controlled combustion. Note the near matched expansion pressure confirms that fuel energy is matched for all cases. ...............................................................................................................156 Figure 72 Indicated cylinder pressure and AHRR of E-98 EGR sweep at 9 (bar) IMEP, for the tested conditions shown in Table 16. ..........................................160 Figure 73 Combustion performance and timing with E-98/diesel, gasoline/diesel and gasoline/gasoline + 1.75% DTBP. Note that ISFC is presented on a gasoline equivalent basis ................................................................................................161

xiv

List of Tables Table 1 3401 SCOTE geometry..........................................................................48 Table 2 thermocouple instrumentation location and purpose, adapted from Staples [50] .........................................................................................................50 Table 3 Laboratory Pressure Transducers, adapted from Hardy [13] .................51 Table 4 Low Pressure Fuel System Hardware....................................................53 Table 5 Port fuel injector specifications...............................................................54 Table 6 Common Rail System Hardware............................................................55 Table 7 Common rail injector nozzle specifications ............................................58 Table 8 Gasoline Fuel Properties........................................................................62 Table 9 Diesel Fuel Properties............................................................................63 Table 10 Emissions Analyzers ............................................................................65 Table 11 Experimental engine operating conditions ...........................................75 Table 12 Experimental emissions measurements ..............................................77 Table 13 Direct Injection Operating Conditions.................................................108 Table 14 6 (bar) IMEP baseline operating conditions .......................................114 Table 15 6 (bar) IMEP DTBP percentages fueling ............................................118 Table 16 9 (bar) net IMEP baseline operating conditions .................................123 Table 17 Gross cycle performance comparison of Diesel and 3.5% DTBP 9 (bar) operation with matched conditions....................................................................128 Table 18 Mechanical work breakdown for the fuels shown in Figure 73. ..........162

xv

Nomenclature A AHRR ASI ATDC C CA CAD CAN CH2O CHO CH3CHOH CH3CH2·OH CH3CH·CH5H11 CN CO COV CO2 DEF DI DOC DTBP DPF ECM EID` EGR EGT EPA ERC FFT FSN FTIR g GM GUI H.O. H2O H2O2 HC HCCI HEUI HFID HMN HO2 hr

Amps Apparent Heat Release Rate After Start of Injection After Top Dead Center Degrees Celsius Crank Angle (after top-dead-center) Crank Angle Degrees Controller Area Network Formaldehyde Formyl radical Ethoxy 1-Hydroxy-1-Ethyl Heptyl Radical Cetane Number Carbon Monoxide Coefficient of Variation Carbon Dioxide Diesel Exhaust Fluid Direct Injection Diesel Oxidation Catalyst Di-tert-butyl Peroxide Diesel Particulate trap Electronic Control Module Ignition Delay Exhaust Gas Recirculation Exhaust Gas Temperature Environmental Protection Agency Engine Research Center Fast Fourier Transform Filter Smoke Number Fourier Transform Infrared Gram General Motors Graphical User Interface High Output Water Hydrogen Peroxide Unburned Hydrocarbons Homogeneous Charge Compression Ignition Hydraulic Electronic Unit Injector Heated Flame Ionization Device Heptamethylnonane Hydroperoxyl Hour

16 HRR ICE IDI IMEP IVA K kg kJ kW kPa J L LHV LNT LTC m Mj mm MON NM nm NO NOx NTC NVH ON OH PAH PCCI PD PID PM PRF PRR RCM RON SCOTE SCR SI SOI T TDC VGS Φ 2-EHN

Heat Release Rate Internal Combustion Engine Indirect Injection Indicated Mean Effective Pressure Intake Valve Actuation Degrees Kelvin Kilo-gram Kilo-Joule Kilo-Watt Kilopascal Joule Liter Lower Heating Value Lean NOx Trap Low Temperature Combustion meter Mega Joules millimeter Motor Octane Number Newton Meters Nanometers Nitrogen Oxide Nitrogen Oxides Negative Temperature Coefficient Noise Vibration and Harshness Octane Number Hydroxide Poly Aromatic Hydrocarbon Premixed Charge Compression Ignition Photodiode Proportional Integral Derivative Particulate Matter Primary Reference Fuel Pressure Rise Rate Rapid Compression Machine Research Octane Number Single Cylinder Oil Test Engine Selective Catalytic Reduction Spark Ignition Start of Injection Temperature Top Dead Center Variable Geometry Spray Equivalence Ratio 2-Ethyhexyl Nitrate

17

Chapter 1: Introduction 1.1

Background As an energy conversion device, the internal combustion engine (ICE) has

become invaluable to society. The ability of the ICE to provide economically viable and reliable power for both stationary and mobile applications has resulted in mass production of the ICE in a variety of displacement and design platforms. There are two traditional design platforms for the ICE. The oldest platform is the spark ignition (SI) engine based on the design of Nicolaus Otto, known as the Otto cycle. The other, only slightly newer platform is the diesel engine developed by Rudolph Diesel, and is based on the thermodynamic cycle that bears his name [1]. Regardless of the engine design, the increased dependence of society on the ICE has resulted in a proportional increase in adverse environmental effects from its expanded usage. Primarily, engines have by and large been operated with non-renewable petroleum-based fuel sources. The depletion of the fuel stocks has raised both historical and recent concerns over improving the fuel efficiency of engines. Secondly, because of the widespread usage of engines, the emissions produced have been and continue to be a major area of both concern and improvement. Regardless of the engine design or the type hydrocarbon-based fuel feedstock; the engine is an energy conversion device that converts chemical potential into useful work through combustion. Ideally the products of hydrocarbon combustion would be carbon dioxide (CO2) water (H2O) and nitrogen (N2); unfortunately, this is almost never the case. Typical engine

18 operation produces incomplete combustion products that results in inefficiency and anthropogenic-derived

atmospheric

pollution. Government

mandated

regulation of pollutants such as carbon monoxide (CO), unburned hydrocarbon (HC), oxides of nitrogen (NOx), and particulate matter (PM) have been aggressively increased since implementation, as summarized in reference [2]. To understand how and where improvements in engine efficiency and pollutant emissions can be made, the fundamental mechanisms of operation of each engine platform must be known. 1.1.1

Spark ignition engine fundamentals

The widely popular SI engine relies on combustion to be initiated by an outside means, most commonly a spark. The spark is used to ignite a premixed charge of fuel and air, so-called reactants. Once sparked, combustion ensues from a growing flame front, consuming the reactants and producing products [3]. This method enables stable and controllable combustion, as well as the ability to premix the reactants at stoichiometric ratios to enable the use of three-way catalytic reduction of the engine-out pollutants [1,3]. However, there are efficiency and design disadvantages to the premixed-reactants strategy of the SI engine. Initially there is a macro scale thermodynamic efficiency loss due to the throttling used to control the amount of air in the reactants [1,3]. Secondly, there is an additional combustion efficiency loss resulting from incomplete combustion, including piston-liner crevice flows. Once reactants become trapped in a crevice, the reactants are unable to fully combust, resulting not only in decreased energy conversion, but also increased engine-out emissions of CO and HC that must be

19 treated in the exhaust catalyst. Lastly, there are design limitations on the engine cylinder bore diameter, and cylinder pressure and temperatures. Large bore or high cylinder pressure and temperature SI engines cannot be practically operated due to economics and the commercial availability of the required high octane fuel. The premixed nature of the reactants requires them to resist spontaneous combustion, but still be volatile and rapidly react once combustion is desired. A large cylinder bore, or high pressure and temperature SI engine with insufficient resistance to auto ignition (defined by the fuels octane number (ON)) will result in premature combustion, which can cause efficiency loss and engine damage [1,3]. 1.1.2

Compression Ignition fundamentals

In contrast to the SI engine, the CI engine initiates combustion from the spontaneous ignition of liquid fuel injected into compressed air at high temperature and pressure. The air is un-throttled, but the injected fuel amount is, and can be either injected directly into the cylinder or into a pre-chamber connected to the cylinder. These differences not only reduce throttling losses, but allow the use of increased compression ratios; engine bore diameters, lean stoichiometric operation, and the use of more reactive (and possibly less refined) fuels. The combination of these factors traditionally allows the CI engine to attain higher thermal efficiencies than the SI engine [1,3]. Although there are efficiency and engine bore size advantages of the CI engine, they do not come without penalty. Due to the high temperature combustion environment and locally rich combusting fuel jet of the injected fuel, the emissions produced by CI engines

20 tend to be higher in NOx and PM emissions [1,3]. Furthermore, because the fuel is throttled but the air is not, the CI engine can rarely be operated at stoichiometric ratios where practicalre conversion efficiencies of three-way exhaust catalysts are realized. Although there are technologies available or in place for reducing CI exhaust NOx and PM emissions, such as the use of lean NOx traps (LNT) and selective catalytic reducing (SCR) catalysts for NOx, and diesel particulate filters (DPF) for particulate matter [2], they are costly, are complex to implement and to operate, and can lead to reduced engine efficiency. 1.1.3

Advanced Combustion Strategies

The desire to reduce fuel consumption and cleanly operate the ICE with diesel-like engine efficiency with smooth, low-noise, SI-like operation has helped to motivate many researchers to explore a melding of the two engine designs. These more unified engine designs are often classified as advanced combustion strategies. Such strategies aim to combine the inherent efficiency advantages of the CI engine, with the low noise and post catalyst NOx and PM emission levels of SI engines. A common name given to many of these advanced combustion strategies is low temperature combustion (LTC). Similar to SI engines, LTC strategies employ increased mixing time either from early fuel injection or dillulent control, thus reducing NOx and PM emissions. To maintain high engine efficiency LTC strategies operate with diesel-like compression ratios and are un-throttled. Typically the fuel is auto-ignited, through either a purely kinetically controlled path, as in homogeneous charge compression ignition (HCCI), or through controlling local fuel reactivity, which can be either locally lean or slightly rich.

21 Regardless of the LTC strategy, obtaining low emissions with high efficiency is the elementary concern in the combustion process. Secondly, to be societally accepted, the combustion process must not produce excessive levels of combustion noise. Because the fuel is auto-ignited in LTC strategies, controlling the rate of combustion is difficult. When the combustion is rapid and approaches the limit of volumetric combustion, pressure ringing can be experienced [4], which is undesirable both structurally and audibly. The desire to reduce combustion noise through lower pressure rise rates while operating with an LTC strategy is a difficult challenge, and has presented several technical issues that have hampered LTC engine development. 1.1.4

Research objective

To address the efficiency and emissions issues associated with conventional engine designs, an LTC strategy is studied. The primary objective is to explore areas of engine operation where low emissions and pressure rise are realized, while simultaneously obtaining high engine efficiency. The combination of fuel reactivity and dillulent, control is proposed to increase air-fuel mixing times, and combustion rate is controlled through optimally stratified local fuel reactivity. Incylinder measurements of fuel reactivity controlled combustion process are used to further understand the physical nature of the combustion process, and the effects of non-traditional fuels on combustion are examined for both increasing and decreasing fuel reactivity stratification arrangements.

22

Chapter 2: Literature Review 2.1

Conventional Diesel (Compression Ignition) Combustion

The CI engine initiates combustion from spontaneous ignition of highly reactive liquid fuel that is injected into compression heated and pressurized high temperature

and

density

air.

Combustion

initiates

from

the

chemical

decomposition of fuel, which begins at the time of injection. Liquid fuel vaporization rates can be increased by enhancing air entrainment into sprays and by increasing the fuel surface to volume ratio through fuel droplet and spray breakup. These processes couple the injection event to the combustion event, making combustion phasing and control possible through injection timing. To avoid very high cylinder pressures and incomplete combustion, typical injection timings of a conventional CI engine are in the range of 20 (deg) before or after top dead center (TDC). 2.1.1

Indirect injection systems

Historically, indirect injection (IDI) or pre-chamber systems have been popular due to the increased turbulent mixing provided by the flow into and out of the prechamber, thus enabling relatively low injection pressures to be used, and still provide adequate mixing for stable, clean combustion. The reduced injection pressure requirements of IDI engine designs greatly reduce the amount of engineering and first cost requirements of the engine design process. Although the IDI pre-chamber provides an engineering and first cost economic advantage, the additional turbulence also increases heat transfer, negatively effecting fuel

23 economy and life cycle operating cost [1]. For these reasons, IDI engines are better suited to specific applications where their low first cost and design requirements have helped them to stay competitive over direct injection (DI) engine designs. 2.1.2

Direct injection engines

Unlike IDI engine designs, direct injection (DI) has become the mainstay of the heavy and light duty diesel engine industry. The direct injection design does not use a pre-chamber and thus operates by injecting fuel directly into the cylinder. The injection is commonly directed into a cavity or a bowl machined within the piston, which serves as the combustion chamber, where fuel-air mixing can be accelerated. Because of the relatively quiescent nature of the ambient gas in the DI engine design, the injection event itself must provide sufficient turbulence to ensure fuel breakup and combustion [1]. Most commonly this is accomplished through the use of multiple orifice injector nozzles with high injection pressures, typically on the order of 4 to 12 orifices with 500 to 2000 (bar) delivered pressure, respectively. A comprehensive analysis of the DI combustion process has been provided by Dec [5]. When injected, the fuel exits the nozzle orifices and enters the combustion chamber, carrying significant momentum in the fuel stream. Initially the fuel exits the orifice as a jet, which penetrates into the ambient gas while dissipating its momentum. Momentum dissipation results in the fuel jet being broken into small particles (drops), which enhance the entrainment of the ambient high temperature gas into the spray. The combination of high temperature gas

24 entrainment and increased fuel drop surface-to-volume ratio accelerates fuel vaporization rates. After a certain axial distance down stream from the injector orifice, the fuel momentum and evaporation rates equalize, resulting in a steadystate liquid penetration, or liquid length. The liquid length is a function of the ambient temperature and the density of the surrounding air environment, and the injection nozzle orifice diameter. Kamimoto et al. [6] provided evidence that liquid length does not depend on injection pressure. This lack of pressure dependence can be attributed to a balance of the increased injection velocity increasing fuel breakup and droplet surface to volume ratios thus accelerating evaporation rates to match the increased momentum. Therefore, the only way to increase liquid length penetration is to change the ambient air density or to increase the nozzle orifice diameter. Regardless of the liquid length, the evaporation of the injected fuel can be observed as a negative energy flux in the apparent heat release rate (AHRR) analysis of indicated pressure data. Once the injected, the fuel vaporizes and an equivalence ratio1 distribution develops within the mixed vapor jet. The equivalence ratio is on the order of 2 to 4 through the jet interior while a well-defined equivalence ratio of unity exists at the jet periphery [5]. Like the liquid fuel, the vapor jet continues to dissipate momentum through the combustion chamber. Momentum dissipation causes a significant increase in the jet diameter along the injection axis, and is discussed in detail in [7]. Of specific importance, the oxidation of the mixed jet has been shown to coincide with the onset of combustion. This phase of combustion is

1

defined as the actual air fuel ratio divided by the stoichiometric air fuel ratio

25 commonly referred to the pre-mixed burn, and with conventional CI combustion it is typically only a very small amount of the total energy released. However, with fuels that have long ignition delays and low boiling points like gasoline, the entire combustion event consists of only a pre-mixed burn [8, 9]. Typically pre-mixed burn combustion is not observed with conventional compression ignition fuels, because of both the high fuel reactivity, and the rapid rates of energy release observed during the pre-mixed burn that could lead to engine damage. The overall fuel rich pre-mixed burn ignites rapidly, and thus causes a fast rise in the global cylinder pressure. The rapid onset of combustion in a fuel rich environment is also responsible for the formation of poly aromatic hydrocarbon (PAH) species, or soot precursors. PAH species form from partial oxidization of the locally rich injected fuel jets. So long as the jet is un influenced by the engine geometry, the PAH continue to penetrate into the combustion chamber as the jet’s momentum continues to drive the reacting jet further away from the nozzle orifice. Because of the equivalence ratio gradient within the fuel jet, the center of the jet is over rich and cannot support a flame. Oxidation processes with the jet are thus dependent on mixing to transport the PAH species to the jet periphery where the stoichiometric flame is established. If the injection event persists after the onset of the pre-mixed burn, additional fuel will be supplied to the standing jet plume, forming soot, and lifted flame combustion will persist as a mixing controlled process [10]. The development of the mixing controlled combustion process from just after injection to the steady-state liquid length can be seen in Figure 1.

26

Figure 1 Mixing controlled combustion jet development in crank angle degrees after start of injection (ASI) units, [5]

2.1.3

Compression ignition emissions

The mixing-controlled combustion process is dependent on both the convection of PAH from the jet interior to the peripheral stoichiometric flame, and chemical reaction rates at the flame surface. If the injection persists long enough for these rates to balance, a steady state relation is established and the reacting jet reaches a fully developed structure. This structure is commonly referred to as a standing or lifted flame. In long combustion duration, steady-state stationary application such as a boiler flame, mixing controlled flames remain standing and

27 lifted forming little soot [11]. However, in a non-steady state system like a moving piston engine, mixing controlled flames can be extinguished through piston bowl interactions or expansion cooling of the control volume, thus slowing the reaction rate of the flame and resulting in incomplete PAH oxidation. In such instances of incomplete oxidation, soot emissions can increase dramatically. Although fixed volume combustion systems are less prone to soot emissions than engines, both applications cannot avoid high levels of formation of thermal oxides of nitrogen (NOx). Because of the thin local stoichiometric conditions at the flame location, flame temperatures approach or reach the adiabatic flame temperature. Although this is excellent for energy conversion, it is also ideal for forming high NOx emissions. The well-established and demonstrated Zeldovich mechanism for thermal NOx production explains the chemistry for high NOx emissions that result from temperatures above approximately 1900 K [11, 3].

N 2 + O → NO + N

(1)

N + O2 → NO + O

(2)

N + OH → NO + H

(3)

To represent both NOx and soot on a common figure, the choice of an equivalence ratio (Φ) vs. local temperature plot has become common. This concept was originally presented by Kmimoto [12]. Since then, several researchers have presented various combustion strategies in such coordinates, and Sun [13] has unified some of these concepts, as shown in Figure 2.

28

Figure 2 Equivalence ratio temperature plot (Sun [13])

Soot forms from incomplete oxidation of hydrocarbon species. Primarily soot is a product of locally fuel-rich conditions, therefore, the equivalence ratio must be greater than unity, and the local temperature must be below the oxidation limit. However, if the equivalence ratio is exceptionally low and the local temperature is below the soot formation temperature, soot production will be negligible because the fuel has not yet decomposed into soot-forming species. Unfortunately, this strategy is not practical for engine operation because of the minimal fuel energy release from poor combustion, resulting in exceptionally high HC and CO emissions. Thus, to avoid soot formation, the local temperature should be high and the equivalence ratio should be low. This is a good strategy for eliminating soot, and as shown from the Zeldovich mechanism above, is unfortunately also ideal for the formation of thermal NOx emissions. The red circle in Figure 2, displays that the path for conventional diesel combustion lends itself well for the formation of both soot and NOx emissions, and there is only a small window of soot and NOx-free operation in equivalence ratio temperature space covered by

29 other combustion strategies. The reason for these emissions from conventional diesel combustion can be described by the conceptual model for a diesel jet, as presented by [5]. A complete cartoon of the mixing controlled combustion jet can be seen in Figure 3.

Figure 3 Fully developed mixing controlled combustion jet [5]

Although NOx and soot emissions can be problematic emissions from mixing controlled combustion systems, the heterogeneous nature of the fuel and air, as well as the high combustion temperatures tend to form very low unburned hydrocarbon (HC) and carbon monoxide (CO) emissions. Although these two emissions, like soot, are products of incomplete combustion, they have been shown to be strong functions of crevice flows created during the compression and combustion of homogeneous charge combustion systems [1, 3]. Thus, so long as the fuel is not allowed to penetrate to the cylinder wall through early

30 injections [14], CI emissions of CO and HC will be low. Unlike NOx and soot emission, HC and CO are easily oxidized in the engine exhaust system by inexpensive catalysts such as the diesel oxidation catalysts (DOC). The fundamental combustion path on which conventional diesel relies greatly increases difficulty in reducing soot and NOx emissions to regulated levels such as the EPA 2010 HD mandate. Alternative approaches have been extensively researched to move the combustion phenomena into low NOX and soot regions of the equivalence ratio-temperature space. 2.1.4

Low Temperature Combustion (LTC)

To meet current and future emissions mandates the conventional CI engine has become reliant on exhaust after-treatment devices. These devices require additional engine cost, possible decreased fuel efficiency, or both. Combustion research for the reduction of these emissions in-cylinder has been ongoing. This research has led to the development of several non-traditional or advanced combustion strategies. Many of these strategies have also focused on simultaneously increasing fuel efficiency. These advanced strategies rely on long ignition delays to increase fuel-air mixing, reducing local equivalence ratios or temperature or both [8, 9, 15-19]. To avoid thermal NOx formation and heat transfer loses, the strategies have also focused on decreasing peak combustion temperatures; and thus have been termed low temperature combustion (LTC) strategies. Although the LTC strategies can reduce NOx and soot, while maintaining diesel-like efficiencies, often HC, CO and with some strategies combustion control and pressure rise

31 rate (PRR), are often sacrificed [14, 19]. Ideally, an advanced combustion strategy would minimize the impacts of negative emissions or control effects, while optimizing the NOx, soot, and fuel efficiency gains. 2.1.5

Homogeneous Charge Compression Ignition (HCCI)

A well researched LTC strategy, HCCI, can best be described as a melding of the SI and CI engine combustion strategies. In this strategy, the fuel and air mixture is externally prepared from the cylinder, and is inducted as a homogeneous charge. HCCI operation was initially researched by Onishi et al. [20] in gasoline-fueled two-stroke engines. In 1983, Najt and Foster [21] and Thring [22] demonstrated HCCI operation of four-stroke engines also with diesel fuel. Since these initial studies, several researches have explored HCCI to further understand its operation, where several benefits and limitations have been found, requiring further work. 2.1.6

HCCI operation

As indicated by the HCCI acronym, the strategy is reliant on a well-mixed or homogeneous charge of reactants. Like the modern fuel-injected SI engine, the fuel and air are mixed external to the engine cylinder, often through a port fuel injection/fumigation system. This mixing strategy allows for complete mixing of the reactants external to the cylinder. However, unlike the SI engine, the intake air is un-throttled and thus load is controlled through fuel/equivalence ratio throttling, as in the CI engine. The homogeneous charge is inducted into the cylinder and is compressed by the piston motion. This allows the homogeneous

32 reactant charge to completely distribute itself throughout the cylinder with minimal equivalence ratio and thermal gradients. The reactants rely on chemistry to initiate combustion, and not on an external means, such as a spark plug. This strategy removes the possibility of fuel injection or spark timing to be used for combustion phasing control. Because the homogeneous charge is dependent on chemical kinetics to initiate ignition, the combustion event occurs nearly simultaneously and uniformly throughout the cylinder. Rapid heat release produces rapid pressure rise rates (PRR), which can have adverse mechanical and noise vibration and harshness (NVH) characteristics. Ultimately, as engine load increases, the amount of heat released increases, as well as the subsequent PRR until the limits of the engine are reached. Although some researchers have demonstrated high load HCCI operation [23], typical engine structural or NVH requirements prohibit practical and commercial HCCI operation beyond that of low load. 2.1.7

HCCI research

The inherent dependence of the HCCI engine on chemical kinetics, chemical mechanisms, and reaction rates is highly important. Research on understanding the chemical dependencies on combustion has warranted significant research designed to further understand methods of combustion control. Research on thermal stratification has helped to explain the longer than theoretical, yet still short combustion durations. Experimental Research by Dec et al. [24] and computational modeling by Sjöberg et al. [25] has demonstrated that although HCCI charge preparation is ideally as well-mixed as possible, the combustion

33 event is not. Natural thermal stratification from engine boundary surfaces promotes ignition in hotter areas prior to cooler areas. This natural thermal stratification extends the heat release event, with ignition initiation at the hottest areas of the combustion chamber, and progression to the cooler areas of the chamber. Without such thermal gradients auto ignition would be globally simultaneous, and ultra-high rates of heat release and pressure rise would result. Although the natural stratification is beneficial for extending the heat release event beyond that of theoretical uniform global auto-ignition, the combustion event remains short in duration. The short duration can provide significant pressure ringing and rates of pressure rise [4]. To minimize the effect of excessive pressure, either the combustion duration must be increased, or combustion must be phased appropriately. Research on methods of increasing combustion duration or phasing through compositional [26, 27] or thermal stratification [24] has examined potential means of controlling and extending the operational range of HCCI combustion. Herold et al. [28] has demonstrated that compositional and thermal effects can compete during the ignition process of HCCI operation and careful management of both thermal and reactant composition distributions are needed for successful stratified combustion control. Such methods for combustion phasing control effectively alter the relation between chemical and engine time scales. Fuel chemistry time is “wall clock” time, and has no reference to engine crank angle time. Thus, phasing combustion can be accomplished through matching or shifting chemical times relative to engine times [29].

34 Combustion can be phased by either coinciding the engine speed and the fuel chemical time, or the fuel chemical time with the engine time. Although simple, this approach is not practical in most situations, and is thus not a robust solution to phasing HCCI operation. A more robust approach is to match the chemical timescale to a given engine timescale. Because the fuel timescale is purely kinetically controlled, this requires modifying the kinetic timescale. Modifications through temperature, collision rate, or chemical mechanisms can be employed. For example, increased pressure results in increased collision rates, and increased collision rates accelerate chemical kinetics, and thus leading to earlier combustion phasing. The effects of temperature and pressure can be altered through compression ratio control, intake air heating/cooling, controlled boost, and trapped residual, or combinations thereof. These effects ultimately modify the kinetics of the fuel reactivity. For example, a highly reactive fuel-like n-heptane, exhibits fast fuel kinetic timescales. To phase combustion later to avoid excessive pressure rise rates requires slowing the fuel kinetics by lowering the operating temperature or pressure. The dependence of the combustion event on fuel reactivity is ultimately the defining factor of HCCI combustion phasing [17, 30]. Significant research to understand fuel reactivity effects on HCCI combustion has been conducted. A sample of such research has been presented by Shibata [31, 32], which studied fuel effects on HCCI. His results support other research that shows that more reactive fuels, like normal paraffin fuels, exhibit

35 low temperature heat release [17], with higher low temperature heat release amounts from more auto ignition-prone fuels [33]. Because many fuels demonstrate both low and high temperature heat release, both high temperature and low temperature chemistry need to be studied. Shibata [31] also demonstrated that that the high temperature heat release timing is dependent on both the total amount of heat released and the timing of the low temperature heat release event. When the amount of heat released in the low temperature heat release is high, the temperature history of the fuel-air mixture is increased, and fuel chemistry rates increase. This can be likened to operation without low temperature heat release but with elevated kinetic rates (higher initial fuel reactivity). Secondly, the timing of the low temperature reactions is critical to the phasing of the high temperature heat release event. The reactions inherent in the low temperature chemistry not only release heat, but lead to the formation highly reactive radicals. This chemistry is responsible for the transition between the main or high temperature heat release and the low temperature chemistry leading to thermal ignition. 2.2

Hydrocarbon Oxidation

2.2.1

Fuel Decomposition Steps

Hydrocarbon-based fuels are organic chemical molecules, and their overall oxidation process is an exothermic reaction process. Much research has been done on chemical processes occurring during fuel decomposition, and works involving detailed chemical kinetic theory or fuel ignition processes, can be seen in references [30, 31, 32, 34]. Complete detailed reaction systems often consist

36 of hundreds of chemical species and thousands of individual reactions to fully describe the oxidation of a single hydrocarbon fuel species. Although in-depth, detailed studies of hydrocarbon oxidation are required to investigate fundamental chemistry, the vastness of a detailed analysis is impractical and often unnecessary for a majority of engine research. Thus, several researchers have examined ways to reduce the number of chemical species and chemical reactions necessary to properly describe the fuel oxidation process of internal combustion engines. Through the identification of key chemical species and reactions, the major steps of the fuel oxidation process can be used to study the combustion process more readily. Figure 4 depicts an example that illustrates the relatively small number of key reactions that can result from detailed analysis of hundreds of actual reactions and species by Ando et al. [29].

Figure 4 Contribution matrix of a reduced reaction set leading to a four step hydrocarbon oxidation system [29].

37

Although the overall process is extremely complex, grouping the oxidation into major groups has been very useful in understanding the decomposition process in particular Ando et al. [29] and Westbrook [35] have broken the oxidation process into four steps labeled: fuel series reactions, fuel fragment reactions, the H2O2 loop reactions, and the H2–O2 reactions, see Figure 4. Each step is required to reach a runaway reaction, also called thermal ignition, which leads to complete combustion. Of the four steps, the first and most basic step consists of the fuel series reactions. During this step, large fuel species react by breaking up into smaller fuel species, known as fuel fragments. The fuel fragments are not strictly hydrogen-and oxygen-based molecules; the fragments can also be oxygenated species. The second step of the oxidation process is where reactions of the fuel fragments take place. During this step, the fuel fragments react to form new smaller fragments and initiate the formation of radicals. Radicals are highly reactive and accelerate the oxidation process to form species such as formaldehyde, which allow transition into the H2O2 loop. In the H2O2 loop, highly reactive species such as OH form in abundance, and exothermic heat release occurs. The main reaction pathways of the H2O2 loop mechanism are depicted in Figure 5, where OH branching is seen as the main highly reactive by-product of H2O2 chemistry.

38

Figure 5 H2O2 Loop mechanism as depicted in [28] The positive amount of heat released in the H2O 2 loop through H2O and other radical species formation accelerates the chemical kinetics to engine time scale and ensures that combustion occurs and is phased accordingly [29, 31]. Once the H2O2 loop has accumulated both a sufficient number of radicals and there has been a sufficient rise in temperature, the oxidation transitions into the last step, thermal ignition. Thermal ignition, also known as the H2–O2 system, is where rapid highly irreversible chemical reactions and thus energy release occurs. This last step is where a majority of the chemical energy is released. 2.3

Cetane Improvers

The cetane number (CN) is a measure of the ignition quality of a fuel. The definition of the cetane number is based on the behavior of the fuel relative to that of two reference fuels, n-hexadecane (CN=100) and heptamethylnonane (HMN) (CN=15). Where the CN is defined as [3]

CN = (% n _ hexadecane)+ (.15 *% HMN)

(4)

39 Although all fuels have a CN, fuels that resist auto ignition have lower CN (~15 typical of gasoline) while fuels that readily ignite have higher CN (~46 typical of US diesel). The CN of the fuel depends on the fuel composition; the more reactive the components, the higher the CN of the fuel. Using this relation, the CN of a fuel can be synthetically modified through the addition or subtraction of highly reactive fuel species. For example, the combustion behavior of HCCI engines has been characterized by an “HCCI index” by Shibata [31]. This relation expands on the octane (inverse of CN) relation developed by Kalghatgi [30], where Shibata’s modified relation accounts not only for the motor octane number (MON), but also the chemical bond structure of the constituent fuel species. The relation developed by Kalghatgi [30] can be seen in Equation 5, where RON is the research Octane number, MON is the motor octane number, K is a constant that is dependent on temperature, pressure, and trapped residual, and S is the “sensitivity” of the fuel (chemical species dependent).

OI = (1-K) RON + K MON = RON - K·S

(5)

The modified relation by Shibata in equation 6 accounts for the individual species reactivity instead of the lumped parameter Kalghatgi classified as S.

HCCI Index = m·MON + a[nP] + b[iP] + c[O] + d[A] + e[OX] + Y

(6)

40 Where m, a, b, c, d, e and Y are empirical constants, nP is the concentration of normal paraffins in the fuel, iP is the concentration of isomeric paraffins in the fuel, O is the concentration of olefins in the fuel, A is the concentration of aromatics in the fuel, OX is the concentration of oxygenates in the fuel, where the concentrations are in percent volume. If a decrease in the HCCI index and thus increased fuel reactivity (CN) is desired, it can be accomplished through addition of fuel species that have easily broken bonds. Although normal paraffins have such qualities, chemical species with significantly weaker bond structures do exist, and are often referred to as CN improvers. The weaker chemical structure of CN improvers often contain nitrate or peroxide groups, which have easily broken bonds. Two common CN improvers are 2-ethylhexyl nitrate (2-EHN) and di-tert butyl peroxide (DTBP). Only a small percentage of either of these chemicals is required to achieve significant increase in cetane number. Using the correlations developed by Thompson [36] for DTBP and 2-EHN addition to diesel fuel, EPA studies [37] on the benefits of cetane improvers has demonstrated the trends of Figure 6.

41

Figure 6 EPA study of cetane improvement to diesel fuel [37] using the correlations from [36]

Although 2-EHN is a more effective CN improver for diesel fuel, opposite trends are seen with rapid compression machine measurements with gasoline-like fuels [38]. This trend is explained by the fact that both 2-EHN and DTBP decomposition initiates with the production of alkoxy radicals, and the initial dissociation energies for each respective CN improving additive are similar. The decreased ignition delay with DTBP and gasoline-like fuels results from the dissociation of DTBP forming twice the number of alkoxy radicals as 2-EHN. The nearly doubled radical density results in approximately double the ignition delay improvement over the base fuel. Figure 7 presents this finding from rapid compression machine (RCM) tests using primary reference fuel (PRF) 90 both with and without the respective additive.

42

Figure 7 Rapid Compression Machine (RCM) ignition characteristics of PRF blends with 2% by volume of 2-EHN and DTBP. Note that the DTBP is more effective with a PRF representative of gasoline than 2-EHN, opposite to the behavior with diesel fuels [38] Similar to n-heptane, either DTBP or 2-EHN accelerates the early chemistry and radical formation. However, the additives decompose more readily than nheptane due to the lower activation energy pathways. The low energy pathways are due to the weak molecular bonds within the nitrate and peroxide compounds, and are responsible for lowering ignition delay. Supporting the rapid compression machine findings of Tanaka [38], results of DTBP ignition enhancement by Eng et al. [39] has demonstrated successful HCCI operation with DTBP addition to PRF blends. Although both DTBP and 2-EHN have been shown to reduce ignition delays both in theory and practice, the nitrate group of 2-EHN has demonstrated

43 increased NOx emissions in low temperature combustion regimes [40, 41]. For these reasons, LTC operation with cetane improvement might be better achieved via cetane improvement with peroxide groups. The peroxide group of DTBP has a low bond strength and has been shown to have the chemical pathway depicted in Figure 8 [42].

Figure 8 Chemical mechanism of DTBP as proposed by [42], note the scission of the peroxide bond occurs first. The overall effect that CN improvers have on the overall fuel oxidation has been shown to be insignificant. Higgens et al. [43] have shown that the addition of less than 1% (by mass or volume) of CN improver to diesel fuel dramatically shortens the ignition delay of the fuel. However, their findings also noted that although the ignition delay decreased with the addition of CN improver, the combustion characteristics of the base fuel remained unaltered. Similar findings by Tanka et al. [38] as seen in Figure 9 have supported the results of Higgens et al. [43] where the CN improvers significantly decrease ignition delay, but had a minimal effect on the total burn duration, as shown in Figure 10.

44

Figure 9 Ignition delay and burn rate of various cetane improves vs. initial temperature. Note that cetane improvers have little to no effect on burn duration Tanka el al. [38]

Figure 10 Combustion luminosity of high and low cetane diesel and low cetane diesel with 2-EHN cetane improver. Note that the cetane improver has no effect on luminosity trends, it only advances the time of initial luminosity [43]

Similar ignition delay reductions from CN improvers with conventional liquid fuels have been observed with gaseous fuels, where slightly higher percentage additions of CN improver to low CN fuels, like LPG [44], have resulted in diesel

45 fuel-like behavior. These studies demonstrate that reactivity enhancement can be achieved with addition of CN improvers to base fuels with natural CN as low as nearly zero, and that there are no adverse main heat release effects from the addition. 2.4

Fuel reactivity stratification for extended duration fully premixed combustion

An alternative strategy that researchers have investigated for increasing the load range of HCCI has been through fuel blending external to the engine. Shabata [31] has demonstrated that blending of various reference fuels external to the engine can result in complex fuel chemistry with either a broadening or narrowing of the low and high temperature heat release rates. Building on this work Kalghatgi [30, 34] and Bessonate et al. [23] studied fuel reactivity and load limits of HCCI. Although Bessonate was able to achieve loads of 16 (bar) IMEP with HCCI operation, significant pressure rise rates (up to 30 [bar/deg]) were also obtained. Both Kalghatgi and Bessonate demonstrate that at higher load conditions the fuel reactivity for optimal combustion phasing and lower PRR is that of a low cetane number fuel. Although their results are promising, commercially available low cetane fuels are not standard, and the fuel reactivity also changes according to climate (winter/summer), and market conditions. To avoid the need for a single fuel reactivity, similar research was performed by Inagaki et al. [45] where the blending of multiple reference fuels found to extend the heat release rate event. However, in that study, the fuels were internally blended within the combustion

46 chamber through the use of separate fuel delivery systems. The results of Inagaki et al. [45] are very promising, and although the fuel delivery strategy is not true HCCI, it is near-fully premixed. This research has demonstrates that through the use of appropriate fuels with proper mixing, a significant increase in load with low-pressure rise rates can be realized while operating in a fully premixed combustion regime. Recent improvements to the work of Inagaki et al. [45], have been demonstrated by Kokjohn computationally [46], and by Hanson experimentally [47, 48]. Unlike the work by Inagaki et al. [45], Hanson used commercial pump fuels instead of reference fuels. The fuels selected were US pump #2 ULSD diesel and Premium grade pump gasoline, were the fuel properties can be seen in Table 8 and Appendix B. Both Hanson and Kokjohn have demonstrated that through the use of multiple optimally-timed direct injections of diesel fuel in combination with port injection of gasoline the resulting combustion can lead to operation with NOx and soot emissions below EPA 2010 HD limits realized incylinder, while simultaneously maintaining low pressure rise rates at mid/high load operation, and with thermal efficiencies up to 53%. This work is different from the work of Inagaki et al. [45] because optimally-timed multiple injections can be used to realize significant CO and HC emission reductions, thus resulting in approximately a 10 percent increase in thermal efficiency over Inagaki’s work. Although these results are exceptional, the strategy requires the use of two separate fuel stocks, thus requiring two fuel systems for operation in an application. Both fuel tanks would require regular refueling, and could result in

47 decreased market acceptance due to the requirement to refuel two systems. An alternative strategy would be use a single fuel stock with the same delivery system as that proposed by Hanson et al. [47, 48], but with the introduction of a small concentration of a highly reactive chemical to the direct injection fuel, as proposed by Reitz et al. [49]. This strategy would, in-turn establish a similar reactivity gradient from a single fuel stock, thus requiring only a single fuel to be refueled by the end user. The system could be properly sized to allow the reactivity enhancer to be replenished at oil change-like intervals while necessitating a much smaller tank size, thereby aiding in packaging and use of the system. Similar systems have been designed and implemented into current market products for the storage and injection of diesel exhaust fluid (DEF) for use with selective catalytic reduction (SCR) catalysts for NOx aftertreatment control.

48

Chapter 3: Experimental design 3.1

Engine

The present engine experiments were conducted with a heavy-duty 2.44 (L) Caterpillar 3401 Single Cylinder Oil Test Engine (SCOTE) at the University of Wisconsin Madison Engine Research Center. The 3401 is a single cylinder version of the commercially produced Caterpillar 3406E six cylinder diesel engine. Unlike the 3406E, the SCOTE configuration is rated at 62 [kW] (83 [hp]) at a speed of 1800 revolutions per minute. The specific engine geometry and configuration are given in Table 1 and the lab setup is given in Figure 11.

Displacement Bore x Stroke

Table 1 3401 SCOTE geometry 2.44 (L) 13.72 x 16.51cm

Connecting Rod Length

21.16 cm

Squish Height

0.157 cm

Geometric Compression Ratio

16.1:1

Effective Compression Ratio

15.1:1

Swirl Ratio

0.7

Piston Bowl Design

Stock (Mexican Hat)

Number of Valves

4

Intake Valve Opening

-335º ATDC Firing

Intake Valve Closing

-143 º ATDC Firing

Exhaust Valve Opening

130º ATDC Firing

Exhaust valve Closing Engine Cooling system

-355º ATDC Firing

Engine Oiling System

Engine Driven Pump, with Wet Liner External Electrically Driven Oil Pump, With Rotella T15W-40 with Advance Soot Control Oil

49

Figure 11 Diagram of the engine lab

The engine was instrumented with type K thermocouples manufactured by Omega Engineering inc. The Thermocouples were used to verify proper steady state operation of the engine while collecting data. The locations and purpose of each thermocouple can be found in Table 2.

50 Table 2 thermocouple instrumentation location and purpose, adapted from Staples [50] Thermocouple Location Significance of Measurement Used in heat release calculation, for Engine Coolant (in and out) control feedback to prevent overheating, for steady state Used for control feedback to prevent Engine Oil (in and out) overheating, for steady state equilibration indicator Used in heat release calculation, to Exhaust Port and Exhaust Surge prevent overheating, for steady state Tank equilibration indicator Used for control feedback to prevent EGR Heat Exchanger Exit EGR pump overheating and maintain EGR temperature EGR Pump Exit Used for test consistency Used for control feedback to maintain Fuel Supply Temperature fuel temperature Used for diagnostics and test Fuel Return consistency Used for control feedback to maintain Intake Air intake air temperature, for defining engine operating condition Critical Flow Orifice Upstream Used in intake air flow rate calculation Temperature Used for control feedback to maintain Common Rail Fuel Temperature common rail pump inlet temperature of 31° °C

The laboratory was also instrumented with a variety of pressure transducers. The pressure transducers were used to monitor operating conditions, and to verify that steady state engine operation had been achieved. Table 3 displays the location and model of each transducer or analog gauge. The intake and exhaust surge tank pressure transducers were electronically connected to PID controllers manufactured by LOVE industries model 16111 for control logic of the desired pressure. The common rail pressure transducer was connected to a Mototron

51 ECU model ADEM3v.8 for control of rail pressure, and further discussion of this system can be seen in section 3.1.5. Table 3 Laboratory Pressure Transducers, adapted from Hardy [13] Measurement Method

Significance of Instrument Manufacturer/Model Measurement

Cylinder Pressure

Reference Transducer

Kistler 6067C1, amplified by Kistler 5010 Dual Mode Charge Amplifier

Combustion Diagnostics and Heat Release Calculation

Critical Flow Orifice Upstream Pressure

Absolute Transducer

Omega PX203-100 A5V

Intake Air Flow Calculation

Exhaust Surge Tank

Absolute Transducer

Omega PX215-200 AI

Intake Surge Tank

Absolute Transducer

Omega PX215-200 AI

Oil System

Reference Gauge

K.O. Trerice 0-100 [PSIA]

Common Rail

Absolute Transducer

Bosch CR/DRV-P S K/20S 0 281 002 507

Common Rail Pump Inlet

Reference Gauge

Port Injector

Reference Gauge

Pressure Measurement

3.2

Injection Systems

3.2.1

Low Pressure Injection Hardware and Plumbing

Control Feedback and Set Engine Condition Control Feedback and Set Engine Condition Verify Proper Operation Determine Delivered Injector Fuel Pressure Determine Pump Inlet Fuel Pressure Determine Delivered Injector Fuel Pressure

Fuel reactivity controlled combustion engine experiments were performed using port fuel injection of high volatility fuel and early direct injection of more reactive fuel. The port and direct injection fuel systems are completely independent, enabling fully flexible control of each. The low-pressure injection system, as

52 described by Hruby [51], was retained for the experiments. A schematic of the system can be seen in Figure 12 and Table 4 displays the system hardware components.

Figure 12 low pressure fuel system adapted from Hanson [48]

53

Table 4 Low Pressure Fuel System Hardware Pump Hydra Cell F20GASTNEMG Pump Motor Dayton ½ HP 2M168C Motor Controller Dayton SCR Control Boost Pump Unkown 130 psi pump model E8228 Pressure Regulator Parker 637B-3-1/8 Endress-Hauser Promass A Fuel Flow Meter (SN#3N547253)

To convert the engine to port fuel injection, the intake elbow was modified. A custom injector clamp was designed and installed, and the modified and installed assembly can be seen in Figure 13.

Figure 13 installed port injector in intake elbow

The port fuel injector used was a common spark-ignition engine style commercially available automotive injector. To size the injector for the SCOTE an

54 aftermarket performance injector was selected, and its specifications can be seen in Table 5.

Manufacturer Injector Style

Table 5 Port fuel injector specifications RC Engineering Lucas peak and hold

Peak and Cold Current

2 (pull in), 4(hold)

Steady flow rate @ 3 bar

750 cc/min (at 3 bar)

Included Spray Angle

15°

Fuel Pressure

5.17 bar

3.2.2

High Pressure Injection Hardware and Plumbing To deliver high pressure fuel to the engine, a common rail system was

used. The system was originally implemented for use with the variable geometry spray (VGS) experiments performed by Weninger [52]. A schematic of the system can be seen in Figure 14, with the system hardware listed in Table 6.

Figure 14 Common Rail Fuel Delivery System adapted from Hanson [48]

55 Table 6 Common Rail System Hardware Bosch CR/CP3S3/R110/30-789S Common Rail Pump (0 445 020 030) Pump Motor Emerson AD77 5 [hp] AC motor Motor Controller Baldor AC VFD 1D11205 - EB Parker Polyflex SN: 45-13370 Flexible Line (pump to rail) Hose Type: 4005ST 432307-001, 0.177” Bosch A004 153 67 28 Common Rail RD 000 2 Heat Exchanger Alfa Laval CB14 -14 Boost Pump Holley Red Electric Temperature Controller Omega CN8500 PID Pump Coupler Lovejoy L100 11499; 11518 Water Separator NAPA 3123 Primary Rail Pressure Control Valve Bosch 0281002507 Fuel Flow Meter Pierburg instruments 10E

Although the common rail system was originally implemented by Weninger [52], for the variable geometry spray (VGS) experiments described by Sun [sun thesis]; for this research, modifications were performed to adapt the common rail system into the engine’s HEUI head, as described by Liechty [53] and modified for use of the 315B injector by Staples [50]. The modification consisted of the design and fabrication of an insert and clamp system to install a centrally mounted injector in place of the 315B HEUI injector. The insert was designed to mimic the exterior profile, o-ring dimensions, and brass crush insert of the 315B injector. The insert was also designed to be compatible with Bosch or Denso Common rail injectors, and Bosch or Delphi DGI injectors. An exploded view of the entire adaptor assembly with a Bosch common rail can be seen in Figure 15 and all technical drawings for the assembly can be found in Appendix A.

56

Figure 15 Common rail adaptor insert assembly for use with HEUI 315 inserts

Although the insert is designed to mimic the exterior profile of the 315 B injector, the stock Caterpillar part number 315B o-rings are toleranced to fit extremely tight, by making the insert extremely difficult to remove. To alleviate this problem, two Parker o-rings were installed on the top two o-ring lands, and a single Parker o-ring was installed on the bottom land. The installed assembly and Bosch common rail in the HEUI head can be seen in Figure 16.

57

Figure 16 Bosch Common rail injector installed into HEUI 315B head with custom insert (purple) and clamps (orange and green)

Although the insert is compatible with several different injectors, a Bosch common rail system was selected for this research. The injector body is a standard OEM part on the commercially available General Motors (GM)/Fiat 1.9 (L) diesel engine used extensively throughout the Engine Research Center (ERC), with injector part number 1206097959002. The GM/Fiat engine is

58 approximately 0.5 (L) per cylinder, while the SCOTE is 2.44 (L) for a single cylinder. The approximatly 5 times per cylinder displacement increase necessitated the need for a larger injector nozzle flow rate. An appropriate injector nozzle was selected from a commercially available 8.3 (L) high output (H.O.) Cummins QSC-600 marine engine, and installed on the GM/Fiat injector body. The injector body and nozzle were verified to not exhibit choking by rate shaping experiments performed by Tess [54]. Details of the larger flow rate nozzle and general information and dimensions can be found in Table 7.

Table 7 Common rail injector nozzle specifications Manufacturer Bosch Commercial Engine Design Cummins 8.3 L QSCH.O.-600 Number of holes 6 Hole diameter 250 µm Included spray angle 145° K factor 0 Hydroground yes Sack Style Mini-sack Cummins part number 4993482

3.2.3

Electronic Injection Control

With all necessary the hardware for implementing the common rail and port fuel injectors in the laboratory, injection signal and fuel pressure control were required to fuel the engine. Previously, Weninnger [52] had implemented a Caterpillar ADAM v3 electronic control module (ECM) with increased current capacity for controlling the injection signal timing, current, and duration of the higher than HEUI current requirements of the Bosch injector (20 amp pull in, 12.5 amp holding currents). The modified ECM can only send a signal to an injector and cannot acquire engine sensor signals; and is thus called the slave ECM.

59 Because of these limitations, an additional ECM was necessary to acquire engine sensor signals and to trigger the slave ECM to fire the injector. The additional ECM is called the master ECM. The master ECM not only triggers the slave ECM to inject, but it also can fire its own injector; and is thus the ECM where user-defined injection commands are inputted from the Caterpillar CADETwin software installed on a desktop PC. Although this master-slave ECM system can actuate two different injectors within the same engine cycle, there are injection windows that prevent fully flexible injection timings. The injection windows can be seen in Figure 17.

Main and Close-Coupled (cc) Injection Window

Pilot Injection Window

pilot

cc_pilot

-40 ATDC

main

Post Injection Window

cc_post

post

40 ATDC

Figure 17 injection windows of Cadetwin software For this research an injection strategy of two injection within the pilot injection window and a single injection with the main injection window was desired. As seen from Figure 17, there is only a single pilot injection allowed by the masterslave configuration. To solve this issue, an additional ECM without injection windows was installed in parallel. This ECM was the intake Valve Actuator (IVA)

60 ECM used by Nevin [55] for his variable intake valve closure timing studies. The IVA ECM is also based on Caterpillar’s ADAM v3 ECM, but it does not have injection window limitations. The IVA ECM is not modified for higher current output and thus must be used with injectors such as the Stock HEUI or the present low pressure port injector, which required less than 15 amp pullin current output. 3.2.4

Electronic Common Rail Pressure Control

Weininger [52] developed a fuel pressure control strategy for the common rail system very similar to that of Lee [56]. The system used a MotoTron standalone ECM with MotoHawk software for the control architecture, and MotoTune software was used to flash the ECM with calibrations and user-defined operating parameters. MotoTune uses a graphical user interface (GUI) display for user input. The entire system software suite is unlocked by use of a key “dongle” that is placed in the USB port of the lab computer. The dongle pings a server in order to validate the use of MotoTune. The dongle USB port is user definable, and is currently set to PCM-2 (lower left USB port). The entire system communicates over controller area network (CAN) bus. A CAN-adapter (Kvaser Leaf Light HS) connects the computer via PCM-1 (upper left USB port) to the MotoTron junction block, leading to the ECM. The control strategy utilizes a Proportional Integral Differential (PID) control system. The signal from the Bosch common rail pressure transducer is input to the PID, with the linear fit calibration of Figure 18. The PID outputs a signal to actuate a pulse width modulated (PWM) normally open Bosch common

61 rail pressure bleed solenoid. Note: that system designed by Weininger [52] uses only a single pressure control solenoid on the rail, where the stock Bosch system uses a solenoid on both the common rail pump for coarse pressure adjustment and the rail solenoid for fine pressure adjustments.

Figure 18 Common Rail Pressure Transducer Calibration The common rail pressure transducer calibration and PID settings used by this system can be seen in the MotoTune user interface. The calibration is saved as cal/VGS_2, and the display is saved as display/VGS_2 on the laboratory PC. The maximum safe pressure obtainable by this system is 1500 bar, and should never be exceeded. The MotoTune settings for proper system operation can be seen in Figure 19.

62

Figure 19 MotoTune VGS User Interface 3.3

Fuel Properties

Commercially available 91 pump octane number (PON) gasoline and #2 ULSD were used for all tests. Tables 8 and 9 display the respective independent lab analysis of the fuel properties. Table 8 Gasoline Fuel Properties Distillation Curve ASTM D86 Initial Boiling Point 38.9 °C Temperature 10% evaporated 69.4 °C Temperature 50% evaporated 105.0 °C Temperature 90% evaporated 160.6 °C Final Boiling Point 215.6 °C Lower Heating Value 43.2 MJ/kg MON 87.8 RON 95.6 (R+M)/2 (PON) 91.6 Ethanol 0% H/C ratio 1.88 Specific Gravity (@ 15.6°C) 0.737 Sulfur (ASTM D5453) 4.6 ppm

63

Table 9 Diesel Fuel Properties API Gravity (@ 15.5 °C) 33.9 °C Viscosity (@ 40 °C) 2.71 °C Surface Tension (@ 25 °C) 30.0 dyne/cm Density (@ 25 °C) 0.8478 g/cm3 Lower Heating Value 42.526 MJ/kg Cetane Number 46.1 H/C ratio 1.74

Note that gasoline does not normally have any lubricity additives. Thus, when using gasoline in the diessel common rail injection system, 500 ppm of the lubricity additive Infinium 655 was added to the fuel. The properties of Infinum 655 can be seen in Appendix B. 3.4

Data Acquisition and Laboratory Hardware

All important laboratory parameters were recorded with two separate computer software programs, both developed by Theil [57]. The first program, named EnDAQ, was used for monitoring real time engine operation by displaying fuel flow, cylinder pressure and pressure rise rate, (PRR) combustion coefficient of variation (COV), heat release rate and combustion durations, as well as engineout emissions. EnDAQ was also used to save all engine cylinder pressure and heat release rate parameters, by averaging 499 cycles (the most the software allowed). The second program, AutoOpt, was used to calculate and save all emissions and fuel consumption on a brake specific basis. Both EnDAQ and AutoOpt were installed on the laboratory PC, which was interfaced with all engine sensors through a National Instruments model BNC2090 analog interface box, ass shown in Figure 20. The BNC-2090 box was

64 connected to the PC computer by a National Instruments PCI-6071E data acquisition card with the configuration seen in Figure 21.

Figure 20 Analog signal schematic, as documented by Hardy [13]

Figure 21 Serial cable communication schematic, as documented by Hardy [13]

65 3.5

Engine Out Emissions

All gaseous engine-out emissions were analyzed with a five gas analyzer bench, and particulates were measured with a smoke meter. The analyzers used can be seen in Table 10, and the gaseous emissions sampling system can be seen in Figure 22. All gaseous emissions measurements were performed with the 5 gas emissions bench. Gaseous emissions were averaged for 30 seconds after attaining steadystate for several minutes. PM measurements of filter smoke number (FSN), mass per volume (mg/m^3) and specific emissions (g/kWh) were related with the factory AVL calibration and averaged between 5 samples of a 2 L volume each with paper-saving mode off. Table 10 Emissions Analyzers Emission Species

Instrument

Exhaust NO/NO2

California Analytic Instruments

Exhaust CO

California Analytic Instruments

Exhaust CO2

Siemens

Exhaust Unburned Hydrocarbons

Siemens

Intake CO2

California Analytic Instruments

Particulate Matter

AVL

Measurement Method Heated Chemiluminescence 400S-HCLD Detector Non-Dispersive 3300A Infrared Light Absorption Non-Dispersive 5E-2R Infrared Light Absorption Heated Flame FIDAMAT 5E-im Ionization Detector Non-Dispersive 100 Infrared Light Absorption 415s smoke Filter Paper Opacity meter Model

66

Figure 22 Gaseous emissions sampling system, adapted from Lichety [53] 3.6

Optics

3.6.1

Optical access

To gain optical access to the combustion chamber a modified cylinder head was used that was designed and manufactured by Caterpillar for engine research. The cylinder head was modified to have two optical access points. The optical access points were at Locations A (bowl) and B (squish) shown Figures 23 and 24. The cylinder head technical drawings are presented in Appendix A.

67

optical access Location A optical access Location B

common rail fuel spray

pressure transducer location

shading depicts squish area

Figure 23 Bottom view of optical access

Location B with dummy plug installed

common rail injector fiber to FTIR

Location A with optics installed

common rail fuel spray

Figure 24 Isometric view of optical access

68 Optical access locations A (bowl) and B (squish) are at respective angles of 36 and 37 degrees from the bore centerline, with both locations flush cut to the cylinder head firedeck. As diagrammed in Figures 23 and 24, the field of view from location A was near the center of the bowl, while location B provides access to the squish region near the squish/bowl interface. By using two separate optical access locations, spatial detail of in-cylinder mixing and fuel reactivity stratification could be further studied by the point measurement technique used in this study. 3.6.2

Optical technique and hardware

Optical diagnostics were employed to further understand the ignition process of dual fuel PCCI combustion. The optical diagnostic technique enabled cycleresolved spectra measurements of natural thermal emission. The technique was initially demonstrated by Rein et al. [58 and 59], where measurements of natural thermal emission water spectra were conducted through a modified optical spark plug on a small displacement spark-ignition engine. Since then Rein et al. has expanded the technique for use with the SCOTE. A complete explanation of the measurement technique can be found in Summer et al. [60]. The technique has been adapted for the present research through the implementation of a novel optical instrument that consists of a 3.25 (mm) diameter by 210 (mm) long sapphire rod jacketed in a 9 (mm) outer diameter Kovar housing. The instrument allows infrared wavelength thermal emission from combustion to be transmitted from the engine to an FTIR via a fiber optic cable. The instrument was installed in one of the two available optical access holes in the modified cylinder head, as

69 further described in section 3.2.1. An exploded view of the entire optical instrument and fiber assembly can be seen in Figure 25.

Tightening Nut Sapphire Fiber Assembly Fiber Connector Gun Drilled Kovar Rod

Copper Crush Seal

Sapphire Rod

Figure 25 Optical Instrument The fiber optic cable used was a 550 µm core diameter sulfide fiber (As2S3), which pitched light collected by the optical instrument into a standard Nexus 670 FTIR manufactured by Thermo Fisher Scientific. Two additional photodiodes (PDs) and a beamsplitter were introduced into the FTIR to provide easy access to the relevant interferometric signals. These modifications are not required; the

70 signals could also be obtained from the original PDs in the FTIR, provided one has the means to access these signals without disturbing the interlocks integrated into the FTIR. The beamsplitter (S1 in Figure 26) was placed in the beam path of the HeNe laser in the FTIR. The light reflected by the beamsplitter was focused onto a PD (Thorlabs PDA55, silicon) (B in Figure 26) and was used to monitor FTIR mirror position. The standard interferogram detector in the FTIR was also replaced by a different PD (Kolmar Technologies KISDP-1-J2, InSb) (C in Figure 26) and was amplified with a Stanford Research Systems SR570 current amplifier. All open optical paths and fiber connections were continuously purged with zero-grade nitrogen gas. The engine crankshaft encoder A and Z (A in Figure 26) pulses were continuously recorded. Figure 26 shows the optical setup for the measurement.

Figure 26 Diagram of optical setup

71 The translating mirror in the FTIR was continuously scanned at a frequency, independent of and unrelated to the engine speed. The mirror position, interferometer, and crank shaft encoder signals were simultaneously recorded at 500 kS/s using a LabVIEW-based DAQ device (National Instruments PCI-6123, LabVIEW 8.5). In post processing, a low-pass filter was applied to the interferometer signal to remove the frequency modulation induced by the Michalson interferometer. The original interferometer signal was then divided by this filtered signal to reduce sensitivity to imperfect repeatability of the combustion event. Without this step, noise that was attributed to cycle-by-cycle fluctuations in gas temperature was observed. It is believed that these variations, although small enough to modify the spectral shapes only slightly on a cycle-bycycle basis, were associated with large enough variations in total emitted power to contaminate the interferogram. Next, the data were sorted by engine crank-angle. For each crank-angle, different engine cycles populated different regions of the interferogram, according to the location of the translating mirror. The superposition of the interferometer signals from many engine cycles resulted in a composite interferogram for a given engine crank-angle. Overlapping sections in each interferogram were averaged and gaps in the interferogram were filled using a 1D spline interpolation function to ensure constant spacing between points in the interferogram. The interferogram was then phase-corrected to remove any phase errors caused by imperfect optical alignment or nonlinearities in the detection system.

72 A Flat Top windowing function and zero-padding were then applied before performing a Fast-Fourier-transform (FFT) on the interferogram, yielding a spectrum for each crank-angle. These measured spectra were, in turn, corrected for the spectral response of the system, which was characterized using a 1100°K ceramic emitter (Thermo Scientific EverGlo). Corrections for optical depth of field within the combustion chamber were not considered due to the unknowns of the measured gases. Corrections for piston face and optical path induced emission were not performed because no observed underlying continuum or emission from these sources was observed under hot motored (post fired) or fired operation. Also, corrections for residual trapped EGR and ambient humidity and carbon dioxide emission were not performed because they were minimal, and concentration based measurements were also not performed.

73

Chapter 4: Optical Investigation Results and Discussion 4.1

Optical diagnostics

Using the optical diagnostics discussed in Chapter 3, the technique was implemented to study fuel reactivity stratification PCCI operation. Fuel reactivity stratification has been demonstrated by both Kokjohn [46] and Hanson [47, 48]; to be a successful charge preparation technique that extends of the practical operating limits of low temperature pre-mixed combustion. Kokjohn et al. [46] used multi-dimensional simulations supported by experimental pressure-based heat release rate measurements of Hanson et al. [47] to demonstrate the extension and control of the heat release through optimal reactivity gradient preparation. Although the multi-dimensional simulation predictions were able to show the benefits of reactivity gradients, engine-out experimental measurements are unable to directly validate the model prediction. Thus, optical diagnostics were employed for both computational validation, and for fundamental understanding of the combustion process, as well as for further exploration into the capabilities of the diagnostic technique itself. 4.1.1

Experimental operating conditions

Experimental fuel reactivity stratification operation was conducted by Hanson [48] at engine loads of 11, 9, and 6 (bar) IMEP. These loads are significantly lower than the maximum IMEP of the SCOTE engine (22 (bar) IMEP). To implement the optical diagnostics, the modified cylinder head and optics described in Chapter 3 were installed. According to the design specifications; engine operation with the optical cylinder head installed should be restricted to short

74 duration, and low load, and pressure rise rate conditions [61]. Based on these constraints the optical diagnostics were used with the engine operated at steadystate with the conditions seen in Table 11. Although the operating load is lower that that previously demonstrated, it was confirmed that the combustion regime is identical and thus comparable to the previous higher load operation standard engine tests of Hanson [47, 48].

75 Table 11 Experimental engine operating conditions Nominal IMEP (bar) 4.5

4.1.2

Engine speed (rev/min)

1300

IVC timing (°ATDC)

-143

EGR rate (%)

0

Equivalence ratio (-)

0.24

Intake temperature (°C)

32

Intake pressure (bar)

1.37

Total fuel (mg/cycle)

54.08

Percent gasoline by mass

67 %

Mass of diesel fuel (mg/cycle)

13.16

Mass of gasoline (mg/cycle)

40.47

Diesel injection pressure (bar)

800

Diesel SOI1 (°ATDC)

-58

Diesel SOI2 (°ATDC)

-37

Fraction of diesel fuel in pulse 1

0.62

Diesel injection duration 1 (°CA)

5.07

Diesel injection duration 2 (°CA)

2.34

Experimental and computational agreement baseline

The engine was operated at steady state for each optical measurement, with the use of multiple measurement locations (Location A with view of the piston bowl, and B with view of the squish region). A top view of the two measurement locations can be seen in Figure 27.

76

Location B

Location A Figure 27 Top view of measurement locations and direct injection fuel spray It was not possible to simultaneously collect data at both locations. Thus, to switch measurement locations, engine shut down and restart was necessary for re-installation of the optics. At each measurement location, several tests were performed to ensure measurement and engine operation repeatability. Emissions and cylinder pressure measurements were also recorded for all tests. Table 12 displays sample emissions of one of the tests, and Figure 28 displays the corresponding pressure and AHRR trace of the same test.

77

Table 12 Experimental emissions measurements NOx (g/kW-hr) 0.016 PM (g/kW-hr)

0.002

HC (g/kW-hr)

4.691

CO (g/kW-hr)

15.047

CO2 (exhaust) (g/kW-hr)

464.761

Brake power (kW)

8.12

Brake Torque (NM)

59.66

COV of IMEP (%)

2.14

ISFC g/kW-hr

185.88

PRR bar/deg

4.13

Indicated net thermal efficiency (%)

45

To directly compare the computational predictions to the experimental diagnostic measurements, Kokjohn [66] performed multi-dimensional simulations with the operating conditions of Table 11. The experimental and computational cylinder pressure and heat release were compared to confirm that the model adequately predicts the measured combustion process. As seen in Figure 28, the cylinder pressure and heat release rate of both the experiment and simulation agree well.

78 10 Experiment Simulation

320

6

240

4

160

2

80

0 -20

-15

-10

-5

0

5

10

15

AHRR [J/°]

Pressure [MPa]

8

400

0 20

Crank [°ATDC] Figure 28 Experimental and computational simulation (Kokjohn [66]) of cylinder pressure and AHRR

With good overall agreement between experimental data and computational predictions, the experimental optical diagnostics were implemented and tested at both Locations A (bowl) and B (squish). The diagnostics were implemented to experimentally investigate the fuel decomposition process of in-cylinder blended dual fuel PCCI combustion. Ideally, optical diagnostics would be used to directly measure the consumption of diesel fuel and gasoline, individually. However, due to the bond similarities and combinations present with multi-bond liquid hydrocarbon species, the spectral fingerprints of the individual fuel species are similar. The result is that C-H bonds exhibit broadly-shaped smooth spectra with few distinct observable quantum states. As shown by Klingbiel [62], further difficulty in individual species identification arises from significant overlap in the spectra of ‘diesel-like’ normal alkanes and those of ‘gasoline-like’ iso-alkanes and aromatics. These issues make differentiation of individual fuel species difficult with spectroscopy, and yet more challenging with natural emission spectroscopy.

79 For these reasons, the “real” multi-component fuels with properties previously described in Tables 8 and 9 of Chapter 3 were selected for the experimental testing, while the computations used reference fuels, n-heptane (diesel surrogate) and iso-octane (gasoline surrogate). 4.1.3

Measured Species

As discussed above, experimental measurements of fuel species decomposition would be ideal; the technique utilized for this study was able to discern broadband fuel spectra of the combination of diesel and gasoline, but was unable to discern their individual fuel species. However, the technique was able to measure intermediate species of the fuel decomposition process, but the precise identification of the intermediates was not possible. Past researchers [29, 35] have identified and discussed the importance of several molecules such as the H2O2, HO2, H2O, CO, CH2O, and CHO for transition into the thermal ignition combustion stage. These researchers have also shown that one of the most important intermediate species is H2O2. Although these species would be the most desirable to measure, the experimental design constraints of the diagnostics did not allow for direct measurement of small O-H bond species due to a combination of factors. Preventing their measurement were experimental optical losses, low reaction temperatures and thus low emission, and the shared wavelength domain of these low concentration species relative to those of water and carbon dioxide present from trapped residual EGR and ambient air. However, the technique was able to measure natural thermal emission of aldehydes such as formaldehyde, a combustion intermediate species that forms

80 just prior to entering and remains present through much of the H2O2 loop reactions [29]. The diagnostics also measured natural thermal emission from fuels, water, and carbon dioxide. These measured species were compared to the multi-dimensional CFD simulation results to further clarify the fuel decomposition process. 4.1.4

Broadband Spectra

A comparison of the measured spectra and HITRAN [63] simulated species spectra are plotted in Figure 29. To generate the simulated spectra, the atmospheric conditions of the HITRAN [63] database ground state was compensated for the combustion environment conditions. Compensation consisted of partition-function quantum level considerations for line-strength determination and the incorporation of Doppler and pressure effects to determine line-width. An alternative approach to correcting the HITRAN [63] database could be to use its derivative database for use with high temperature environments, known as HITEMP. Unfortunately, HITEMP currently only depicts a limited number of combustion product species, and not combustion intermediates. For consistency, the HITRAN [63] database was used for all species comparisons. The simulation of HITRAN [63] spectra at temperatures other than 298 K relies on the fundamental equations of absorption spectra. The technique calculates the spectral absorptive coefficient and then corrects the coefficient from absorption to emission response. A detailed discussion and demonstration of the adaptation of the HITRAN [63] database to combustion environment is given by Meyers [65].

81 To calculate the spectral absorptive coefficient of a given species, the experimental pressure, temperature, and species concentration must be known. If one of the three inputs is unknown the correction technique can be iterated through comparing the simulated spectra to the measured spectra, thus determining the one unknown state. To correct the HITRAN [63] simulated spectra, experimentally measured in-cylinder pressure was collected and assumed to be uniform throughout the combustion chamber; however, for a given crank-angle, the local experimental temperature and concentration of the measured species were unknown. To obtain local temperature and species concentrations, the output provided by KIVA CFD simulations preformed by Kokjohn [66] was used. Iterative fitting of the simulated spectra with the measured spectra could have been used to solve for the empirical conditions directly, as demonstrated by Kranendonk et al. [67]. However, because thermal emission levels tend to be very temperature dependant, any cycle-to-cycle variation in temperature would introduce errors into the interferogram and thus make accurate spectral fitting difficult. A similar measurement using absorption spectra rather than emission spectra would minimize this error; unfortunately, the physical constraints regarding the optical access in the engine did not allow for an absorption measurement.

82

-7 deg ADTC H2O + CO2 HITRAN Simulated H2CO HITRAN Simulated

Emission

0.06

0.04

0.02

0.00 Measured Spectra

Emission

0.06

0.04

0.02

0.00 2400

2600

2800

3000

3200

3400

3600

3800

Wavelength (nm)

Figure 29 Relative wavelength locations of Location A (bowl) measured combustion products, intermediates, and reactants relative to HITRAN [63] simulated absorption spectra using CFD predicted species concentrations and temperatures corrected for emission at -7 deg ATDC (units of emission are arbitrary).

Both the cycle-to-cycle variation in the measurement technique, the use of CFD species concentrations and temperatures, and the correction of atmospheric validates spectra of the HITRAN [63] database to combustion environment temperatures and pressure could be responsible for the differences in the line width and line strength measured and simulated spectra. These differences would make direct overlay comparison of the spectra difficult, and thus stacked comparison are presented in Figure 25. Most notable are differences around 2,700 (nm). These differences appear to be more than just line shape differences. The exact reason for the discrepancy existence is

83 unknown but is thought to be due to a combination of optical losses, signal noise, and possible interactions of unknown O-H bonded intermediates. However, regardless of the source the difference, the overall agreement is considered to be excellent for natural emission measurements.

4.1.5

Fuel Spectra

The measured fuel spectra from Locations A (bowl) and B (squish) just prior to combustion are shown in Figure 30. Although non-alkane fuels like aromatics generally exhibit wavelength envelopes shorter and broader than alkanes, the fuels were defined as any emission contributing spectra between 3,330 and 3,500 (nm). -20 deg ATDC 0.015

Location B

(squish) 0.010

Emission

-20 deg ATDC

Location B Emission

0.005

0.000 0.05

Location A

Emission

0.04

(bowl)

0.03

0.0120

0.0408

0.0105

0.0357

0.0090 0.0075

0.0255

0.0060

0.0204

0.0045

0.0153

0.0030

0.0102

0.0015

0.0051

0.0000 3200

0.02

0.0306 Location B Location A

3300

3400

3500

3600

Location A Emission

Wavelength domain defined as "fuel" in this study

0.0000 3700

Wavelength (nm) 0.01 0.00 3200

3300

3400

3500

3600

3700

Wavelength (nm)

Figure 30 Experimental fuel spectra at Locations A (bowl) and B (squish) just prior to the start of combustion (-20 deg ATDC) (units of emission are arbitrary).

84

The spectra for the two locations are similar in overall shape but not identical, suggesting that the port-injected gasoline, direct-injected diesel in-cylinder blending strategy establishes a fuel species (reactivity) gradient throughout the combustion chamber. The increased noise in the spectra of Location B (squish) remains throughout all measured crank-angles, and is attributed to a combination of factors. Primarily, the location has a shallow maximum depth of field (1.97 mm clearance between the piston and head at top dead center), resulting in a shallower optical depth and thus decreased photon emission. The result is a lower signal-to-noise ratio as compared to Location A (bowl). Secondly the physical measurement locations are at different points relative to the fuel spray axis. Location A (bowl) is located between two jet centerlines and Location B (squish) is on a spray centerline (see Figure 27). Although the measurement locations differ with respect to the diesel fuel delivery, the computational results suggest that prior to combustion; swirl-induced bulk gas motion rotates the charge, mixing the peripheral portions of the chamber more. Any variation in the measured species from this charge motion should average out over the duration of the measurement (13,000 fired cycles). 4.1.6

Intermediate species measurements

As previously mentioned, the optical diagnostics were used to measure aldehydes during combustion, but were unable to differentiate other O-H bonded intermediates, such as those involved in H2O2 loop reactions. Measured aldehydes were identified through comparison of the measured spectra to the

85 HITRAN [63] database simulations of formaldehyde. Because the measurement technique acquired broadband spectra, wavelengths between 3,600 and 3,700 (nm) were defined as "formaldehyde", (in reality, the measurement is actually of fuel and several unknown aldehyde/carbonyl species). This definition of formaldehyde was justified by the growth and decay of a bump in the measured spectra at crank-angles after the NTC behavior occurred, where formaldehyde formation is anticipated. Figure 31 graphically represents the rationale used in this assumption.

0.04

1.00

-7 deg ATDC 0.75

0.01 0.00 -25

0.04

0.50

spectra measurement location

-19

Crank Angle (deg) -7 -1 5 11

-13

0.25

0.00 17

23

1.00

Emission

0.03

0.75

0.02 0.01 0.00 3200

0.50

spectra measurement location

0.25

AHRR (kJ/deg)

-20 deg ATDC

Measured Spectra -7 deg ATDC Formaldehyde HITRAN Simulated

0.040

Defined as Aldehydes

0.030

Emission

0.02

AHRR (kJ/deg)

Emission

0.03

0.020 0.010 0.000 3200 3300 3400 3500 3600 3700 3800

Wavelength (nm)

0.00

3350

3500

3650

3800

Wavelength (nm)

Figure 31 Measured Location A (bowl) spectra at -20 (left bottom) and -7 deg ATDC (left top) and measured spectra as compared to HITRAN [63] simulated spectra of formaldehyde (right) (units of emission are arbitrary).

Similar to aldehydes, olefins such as ethylene and other C-H double bonded combustion intermediates will exist during fuel decomposition. Ethylene, for example, exhibits emission spectra from approximately 3100-3400 nm [64], and

86 was shown in the CFD predictions of this study to have similar magnitude and formation trends as formaldehyde. However, ethylene and other olefin species were unable to be experimentally measured due to their low relative emission as compared to formaldehyde. In future, optical losses could be reduced to further examine such lower emitting combustion intermediate species. 4.1.7

Combustion product species measurements

The optical diagnostics were also used to measure the final combustion products of water and carbon dioxide. The wavelength range selected for measurement of these species was between 2,300 and 3,000 (nm), which encompasses the fundamental O-H vibration mode of water at the tested combustion temperatures. If only water emission was desired, the wavelengths of interest would be reduced to 2,400 to 2,650 (nm), where carbon dioxide emission is not also present in the spectra. Figure 32 depicts the emission of water and carbon dioxide in the prescribed wavelengths before and after combustion. Note: the emission scale in the two plots differs by an order of magnitude. Of particular interest, the optical diagnostics were able to measure spectra of combustion products prior to combustion. These species are attributed to trapped internal EGR and ambient air with humidity and CO2. Also, the signal is significantly cleaner and more defined in the spectra acquired after combustion (18 deg ATDC). The jagged features of the spectra at -18 deg ATDC in Figure 32 are attributed to the low concentrations at crank-angles and temperature prior to combustion, combined with optical losses.

87

Location A 18 deg ATDC Spectra contains H2O +CO2 Spectra only contains H2O

Emission

0.20 0.15 0.10 0.05 0.00

Location A -18 deg ATDC

Emission

0.020

Spectra contains H2O +CO2

0.015

Spectra only contains H2O

0.010 0.005 0.000 2400

2500

2600

2700

2800

2900

3000

Wavelength (nm)

Figure 32 Location A (bowl) measured combustion products of water and carbon dioxide emission before (bottom) and after (top) combustion (units of emission are arbitrary). 4.1.8

Comparison Between Computational Simulation and Measurements

A comparison between the measurements to the KIVA CFD computational predictions at Locations A and B was performed. The comparison was performed to investigate the agreement in the relative trends of species formation times between measurements and predictions. Although both KIVA simulations and experimental measurements were recorded on a one crank-angle basis, to show the overall combustion progression at both Locations A (bowl) and B (squish) only select crank-angles are given in Figure 33. The spectra were plotted with Location A (bowl) on the left and B (squish) on the right of each experimental spectra figure, with same scale throughout all presented crank-angles for the particular measured species. Although spectra at Location B (squish) were

88 recorded for all crank-angles, the water and carbon dioxide spectra of early (low emission) crank-angles was at or below the noise floor of the measurement system. From the figure at the earliest shown crank-angle of -20 deg ATDC, the measurements show the presence of fuel, few aldehydes, and very weak water spectra. At this crank-angle the computations also predict little water and formaldehyde. For these reasons, this crank-angle was selected as the baseline crank-angle. The next row in the figure presents measurements and predictions just after NTC, at -13 deg ATDC. At this crank-angle the simulation predicts formaldehyde to start to form at both Locations A (bowl) and B (squish), which agrees with the measurements by the observed bump in the spectra forming between 3,600 and 3,700 (nm). At this crank-angle there is still little change in the water spectra and predictions. By the next depicted crank-angle of -11 deg ATDC, the predictions and measurements show significant formaldehyde formation. The water spectra remain weak (at the noise floor) in Location B (squish) and is possibly slightly higher at Location A (bowl), with both measurements in agreement with predictions, and the discussion of post NTC water behavior in [29, 35]. Although the predicted aldehyde levels should be lower at -11 deg ADTC than -13 deg ADC, the measured fuel and aldehyde spectra appear to be similar in emission magnitude at these crank-angles. The reason for this is because both the measured bulk and KIVA CFD predicted local gas temperature has risen significantly. Increased temperature results in increased overall emission, regardless of weather the emitting species

89 concentrations have decreased. Further discussion and correction of this confounding phenomenon is given in the following section. As combustion progresses into thermal ignition at -7 deg ATDC, the measurements and predictions of aldehyde species remain high, and the onset of increased water emission is seen at Location A (bowl). This is to be expected as the heat release from the experiment and computation show combustion transitioning into thermal ignition at this crank-angle. As combustion nears the peak of the main heat release at -3 deg ATDC, the measurements and predictions show global decay of the aldehyde species (and fuel in the measurements), with greater decay in the bowl as compared to the squish. Also, the water predictions and measurements support this trend through the measured increased water emission and predicted increased mass fraction. Although the early crank-angle water concentration at Location B (squish) was difficult to detect, now the water band spectrum is definable through the emission corresponding to the temperature and species concentration increases. After the peak of main heat release at 3 deg ATDC, both predictions and measurements show low levels of formaldehyde in the bowl and a slightly higher level in the squish region. The water spectra at this crank-angle are stronger than any prior crank-angle, with both measurements and predictions showing combustion products forming in strong abundance. These two species indicate that thermal ignition is complete, and the process is nearing the end of combustion. At the last crank-angle row of 16 deg ATDC combustion has ended, and there is little predicted and measured aldehydes, but there are strong water species

90 predictions and measurements. The qualitative trends of Figure 33 clearly provide experimental evidence supported by the computational predictions that the in-cylinder fuel blending charge preparation causes combustion to initiate in the bowl, where there is higher fuel reactivity, and that reaction progresses to the squish region, where there is lower fuel reactivity.

91

0 .0 0 3200

3400

3600

3800

0 .0 0 0 3200

L o c a tio n A

3600

3800

0 .0 1 5

L o c a tio n B

0 .0 0 3200

3400

3600

3800

0 .0 0 0 3200

W a v e le n g th (n m )

L o c a tio n A

0 .1

0 .0 2400

W a v e le n g th (n m )

3400

3600

0 .0 3

2600

L o c a tio n A

0 .0 3

0 .0 2400

2600

0 .0 0 2400

0 .0 0 0 3200

L o c a tio n B

3400

3600

3800

L o c a tio n A

0 .0 0 0 3200

3400

3600

3800

L o c a tio n B

3400

3600

3800

0 .0 0 0 3200

W a v e le n g t h ( n m )

L o c a tio n A

3600

3800

L o c a tio n B

3400

3600

3800

0 .0 0 0 3200

W a v e le n g t h ( n m )

L o c a tio n A

3400

3600

3800

0 .0 1 5

L o c a tio n B

0 .0 0 3200

3400

3600

3800

W a v e le n g th (n m )

―Spectra

Emission

0 .0 2400

0 .0 0 0 3200

3400

3600

3800

W a v e le n g th (n m )

0 .0 3

L o c a tio n B

0 .0 0 2400

2600

W a v e le n g th (n m )

0 .0 3

L o c a tio n B

0 .0 0 2400

2600

W a v e le n g th (n m )

0 .0 3

2600

L o c a tio n B

0 .0 0 2400

W a v e le n g t h ( n m )

L o c a tio n A

0 .1

Emission

Emission

0 .0 5

2600

L o c a tio n A

0 .1

W a v e le n g t h ( n m )

2600

W a v e le n g th (n m )

W a v e le n g th (n m )

Emission

0 .0 0 3200

0 .0 2400

W a v e le n g t h ( n m ) 0 .0 1 5

2600

L o c a tio n A

0 .1

Emission

Emission

0 .0 5

3400

0 .0 0 2400

W a v e le n g th (n m )

Emission

0 .0 0 3200

L o c a tio n A

0 .0 2400

W a v e le n g th (n m )

0 .0 1 5

2600

W a v e le n g th (n m )

0 .1

Emission

Emission

0 .0 5

0 .0 2400

Emission

0 .0 1 5

W a v e le n g th (n m )

-3 °ATDC

3800

Emission

0 .0 0 3200

3 °ATDC

3600

Emission

Emission

-7 °ATDC

L o c a tio n A

0 .0 5

3400

W a v e le n g th (n m )

L o c a tio n B

Emission

3800

0 .0 3

Emission

3600

2600

W a v e le n g th (n m )

2600

W a v e le n g th (n m ) 0 .0 3

L o c a tio n B

Emission

3400

L o c a tio n A

0 .1

Emission

0 .0 0 3200

W a v e le n g th (n m )

16 °ATDC

L o c a tio n B

Emission

0 .0 1 5

L o c a tio n B

3800

W a v e le n g th (n m )

Emission

Emission

-11 °ATDC

0 .0 5

2600

W a v e le n g th (n m )

W a v e le n g th (n m )

L o c a tio n A

L o c a tio n B

0 .0 0 2400

W a v e le n g th (n m )

0 .1

Emission

Emission

0 .0 5

3400

B

Emission

A

W a v e le n g th (n m )

-13 °ATDC

L o c a tio n B

Water

Emission

0 .0 1 5

Emission

L o c a tio n A

0 .0 5

Experimental Spectra

Emission

B

and CFD Simulation Water

Emission

A

Experimental Fuel Aldehyde Spectra

Emission

-20 °ATDC

CFD Simulation Formaldehyde

0 .0 2400

2600

W a v e le n g t h ( n m )

0 .0 0 2400

2600

W a v e le n g th (n m )

―Spectra

Figure 33 Select crank-angles of measured spectra and predicted species (units of emission are arbitrary).

92 4.2

Spectra Temperature Compensation and Reaction Extent

In order to provide more quantitative results, the emission data were further analyzed. Unlike other optical diagnostics techniques, such as absorption and laser-based diagnostics, natural thermal emission does not use an external light source. Thus, measurements require a detectable number of naturally emitted photons. This lack of external energy input results in a bias of the measurement technique to conditions with higher emission levels, as is the case at elevated gas temperatures. This makes the comparison of measurements acquired at lower and higher gas temperatures difficult. The governing equation for natural emission of a gas is E = ∑ ε i *σ * T n

(7)

i

E is the total radiant emission, εi is the individual species emission, σ is the Stefan-Boltzmann constant, T is absolute temperature, and n is the temperature dependence (local or global depending on wavelength range of interest), where in the limit of a black body emitter, the temperature dependence in is to the fourth power. To compensate, the fuel and aldehyde spectra for temperature dependence, it was assumed that prior to combustion the total radiant emission was only a function on temperature because species and thus εi were assumed to be constant. The temperatures used were those predicated by KIVA CFD simulation. Using this assumption and the KIVA CFD temperatures the total emission value for each non-combusting crank angle was held constant and the

93 power dependence was solved for. As seen in Figure 34, this temperature correction applied at Location A (bowl) resulted in an estimated temperature dependence of the fourth power (bulk temperature derived from the measured pressure also yielded fourth power dependence). Confidence using the chemical mechanism dependent KIVA CFD predicted temperatures at Location A (bowl) and B (squish) comes from the good overall agreement in the bulk cylinder pressure and AHRR trends previously depicted in Figure 28. 8

Integrated fuel emission

Fit Emission = 1.76092E-11·T Integrated fuel emission

4.0712

-18

7.5 -19

7

-20 -22

6.5 -23

6 -25

5.5 680

690

700

710

720

730

Cylinder Temperature [K]

Figure 34 Temperature power (n) dependence calculation where the numbers below each data point correspond to the crank-angle of each integrated fuel emission value (units of emission are arbitrary).

After calculating the temperature power dependence of the measured fuel species spectra, the temperature correction was applied and tested at crankangles ±2 deg of -20 deg ATDC, using the local temperatures predicted by the KIVA simulation. As seen in Figure 35, the results further support that the use of KIVA-CFD-predicted temperatures is reasonable.

94 -22 deg ATDC corrected -20 deg ATDC raw -18 deg ATDC corrected

0.04

Emission

0.03 0.02 0.01 0.00

-22 deg ATDC raw -20 deg ATDC raw -18 deg ATDC raw

0.04

Emission

0.03 0.02 0.01 0.00 3200

3300

3400

3500

3600

3700

3800

Wavelength (nm)

Figure 35 Location A (bowl) raw measured (bottom plot) and temperaturecorrected spectra using CFD simulation temperatures (top plot) (units of emission are arbitrary).

The fourth power temperature correction was assumed to be constant throughout combustion and was applied to all fuel and aldehyde spectra at both Locations A (bowl) and B (squish) from -20 deg ATDC to 20 deg ATDC. Using the fourth power dependence throughout the combustion process eliminates an unknown variable in the total radiant emission equation, and thus equation 7 for total emission becomes a function of species (εi) only. For the temperature correction applied to Locations A (bowl) and B (squish), the resulting spectra are shown in Figures 36 and 37, where the numbers denote the crank-angle of the measured spectra, and the data are presented relative to their baseline value at -20 deg ATDC.

95

0.04

Fuels -20 to -16

Location A

0.03

Emission

-20 to -16

0.02

-14

-14

-12

-12

-10

-10

-8

Aldehydes

-8

-12

-10 -8

0.01

0.00 3200

3300

-6

-6

-4

-4

-2 0

-2 0

3400

3500

3600

3700

3800

Wavelength (nm) Figure 36 Temperature-corrected fuel and aldehyde spectra at Location A (bowl). Only spectra from even crank angles are shown (the units of emission are arbitrary).

Location B Fuels

0.020

Aldehydes

-20 to -11 -10

0.015

Emission

-12 -10

-8

0.010

-8

-8 -6 -6 -4

0.005

0.000 3200

-2

3300

-4

0

-2 0

2

2

4

4

3400

-14 -6

-4

3500

-2

3600

3700

3800

wavelength (nm) Figure 37 Temperature-corrected fuel and aldehyde spectra at Location B (squish). Only spectra from even crank angles are shown, note the delay in combustion as opposed to location A (units of emission are arbitrary).

96 The temperature-corrected spectra displayed in Figures 36 and 37 show the formation of aldehydes occurs during the early stages of the combustion process. To quantify the spectra measurements, the extent of reaction was calculated for the measured spectra. The extent of reaction was defined as; Et − Eo = ξ formation E peak − E o E ∞ − Et = ξ consumption E ∞ − E peak where Et is the total emission at time t and E0 and Epeak represent the emissions at the beginning and ending crank-angles, respectively. The emission is assumed to be a linear combination of all species in the wavelengths of interest, thus spectra of confounded species can be examined. To calculate the extent of reaction, the temperature-corrected measured spectra were integrated between 3,330 and 3,500 (nm) for fuels and 3,600 to 3,700 (nm) for aldehydes. The initial emission crank-angle (E0) is defined as the crank-angle prior to the appearance or disappearance of the species being integrated. E∞ was defined as the last crank-angle where definable spectra (consumption) or the maximum integrated spectra value (formation) of the measured species was obtained. The reaction extent was calculated for both the experimental measurements and the KIVA CFD simulations using the same definitions. 4.2.1

Aldehyde reaction extent comparison

For both Locations A (bowl) and B (squish), the measured and predicted reaction extents of the aldehydes are shown in Figure 38.

97 | |

Experiment Computation

1.0

Consumption

Formation 0.8

Location A

0.6

Reaction Extent (-)

0.4 0.2 0.0 1.0

Consumption

Formation 0.8

Location B

0.6 0.4 0.2 0.0 -20

-18

-16

-14

-12

-10

-8

-6

-4

-2

0

2

4

6

8

10

Crank Angle (deg)

Figure 38 Comparison of reaction extent between experimentally measured aldehydes (solid lines) and KIVA CFD predicted formaldehyde (dashed lines).

The comparison between the experiments and simulations are considered to be fairly reasonable, particularly at Location A, in the piston bowl. However, at Location B (squish) there is a larger discrepancy between the measurement and prediction. One possible reason for the discrepancy could be attributed to the existence of larger reactivity stratification at this location. The measured aldehyde occurs at almost the same time and rate at both Locations A (bowl) and B (squish). At both locations the nearly identical shapes of early aldehyde formation reaction extents suggest that the aldehydes form from the same source, visualizing the early-cycle consumption of the more reactive diesel fuel. This is consistent with the CFD model predictions and is supported from the

98 observation of the aldehydes forming during/after the NTC portion of the experimentally-measured heat release. Figure 38 also shows that there exists a major difference in the mid-life time behavior of the predicted and measured Location B formaldehyde reaction extents. The simulation predictions suggest that additional formaldehyde forms late in the cycle from the consumption of iso-octane. However, the experimental measurements do not support this trend. Also, the measurements do not support the rapid formaldehyde consumption trends of the simulation. The measurements show a slow steady consumption of formaldehyde with only a slightly slower rate to that of Location A (bowl). Although the shapes of the measured and predicted formaldehyde reaction extents slightly differ, the total lifetimes are in good agreement. A possible reason for the observed differences could be the use of multi-component fuels in the experiments and a simplified mix of two singlecomponent pure fuels in the simulation. The computational pure fuels appear to show similar mid-lifetime chemistry, with faster late-cycle kinetics in the squish region as compared to the experiment conducted with real fuels. 4.2.2

Fuel reaction extent comparison

To further investigate the combustion process of in-cylinder blended reactivity stratification, the reaction extent of the fuel was also calculated for both Locations A (bowl) and B (squish). Unlike the experiments, the simulation was capable of distinguishing between each of the two in-cylinder blended fuels. The reaction extent of the individual fuel species and their combination is shown in Figure 39.

99 Experiment | Computation

| 1.0 0.8

n-heptane

Location A

0.6

Reaction Extent (-)

0.4

iso-octane + n-heptane

0.2 iso-octane 0.0 1.0 n-heptane 0.8 0.6

Location B

iso-octane + n-heptane 0.4 0.2

iso-octane 0.0 -20

-18

-16

-14

-12

-10

-8

-6

-4

-2

0

2

4

6

8

10

Crank Angle (deg)

Figure 39 Computational and experimental fuel reaction extent at both Locations A (bowl) and B (squish).

Interestingly, at both locations the measured fuel reaction extents lie in-between the predicted reaction extents. The computationally predicted reaction extent of the combination of iso-octane and n-heptane do match fairly well with both experimentally measured reaction extents, except that they exhibit an S-shaped behavior passing through the measured results. The S-shape stems from the combination of the respective early and late-cycle rapid-reaction extent behavior of n-heptane and iso-octane. Unlike the predicted pure-fuel rapid reaction extents, the linear behavior of the measured reaction extents suggests that the multi-component real fuels behave with a smoother transition from the morereactive to less-reactive species within the fuels.

100 At Location A (bowl) the predicted and measured reaction continually increases, while the less-reactive Location B (squish) trend does not. At Location B (squish), initial reaction starts simultaneously with Location A (bowl) but is followed by a 4-5 CAD delay, after which there is steady reaction until completion. At Location B (squish) there is significantly lower mass of reactive fuel, so when it reacts early in the cycle it is likely to be fully consumed. This provides heat and radicals to the less-reactive fuel, thereby accelerating the kinetic time scale of the less-reactive fuels to match the engine time scale. This trend is not observed in the bowl because its local fuel reactivity blend contains a higher mass percent of more-reactive fuel. This higher local reactivity results in fuel consumption occurring at a steady rate, even as the reaction progresses from more to less-reactive fuel. The result of the radially stratified fuel reactivity is that in areas of low fuel reactivity the reactivity gradient increases the delay from the start of reaction to thermal ignition, whereas there is less delay in locations of higher fuel reactivity. Further supporting evidence is seen in Figure 40 with CFD predicted and measured water reaction extents, and Figure 41 with several CFD predicted species evolutions at Locations A (bowl), B (squish) and the global average. 4.2.3

Water reaction extent comparison

The measured delayed fuel consumption at Location B discussed in section 4.4.2, is also prevalent in examination of the measured water reaction extents. Water reaction extents of CFD predictions and measurements were performed.

101 The respective water reaction extents of both Location A (bowl) and B (squish) can be seen in Figure 40.

1.0 0.8

Experiment Computation

Formation

Location A

0.6

Reaction Extent (-)

0.4 0.2 0.0 1.0 0.8

Formation

Location B

0.6 0.4 0.2 0.0 -20 -18 -16 -14 -12 -10

-8

-6

-4

-2

0

2

4

6

8

10

12

14

16

Crank Angle (deg)

Figure 40 Computational and experimental water reaction extent at both Locations A (bowl) and B (squish).

From Figure 36 the later formation of water in the main heat release period further supports the staged combustion claims of section 4.4.2. Furthermore, the similar early cycle formation of water at both Location A (bowl) and B (squish) also support the aldehyde trends of section 4.4.1. The global production of water at both locations suggests that the reactive fuel fraction globally ignites simultaneously, thus entering the H2O2 loop chemistry phase. The approximate 4 crank angle delay to enter the main heat release agrees with the reaction extent data and supports the staged combustion event hypothesis.

102 4.2.4

Reaction extents and combustion trends

To examine the global significance of the reaction extents calculations and measurements, the reaction extents were compared to the respective heat release event. As can be seen in Figure 41 the CFD simulation (Kokjohn [66]) predicts the experimentally observed combustion delay at Location B (squish). The model predicts that the combustion process is more rapid at Location A (bowl) than both at Location B (squish) and the global average.

-16

-12

-8

-4

0

4

8

12

16

20

0.1

A

Global

0.01

B

1E-3

H2O (-)

-20

1E-4 1E-3

CH2O (-)

Global

1E-4

A

0.01

Global

1E-3

A

B

n-heptane (-)

1E-4

A

Global

1E-3 1E-4

iso-octane (-)

B

1E-5

AHRR

B

1E-5 -20

-16

-12

-8

-4

0

4

8

12

16

20

Crank angle (deg)

Figure 41 CFD species mass-fraction predictions of fuel decomposition globally, and at Locations A (Kokjohn [66]).

The simulation predicts that the early-cycle measured formaldehyde formation in Figure 38 is the result of the consumption of only the reactive fuel. The

103 computations also predict that nearly all of the reactive fuel at Location B (squish) is consumed in the early-cycle reactions. Although the measurements were unable to differentiate the individual fuel reactivates, they supported the predicted trend of early-cycle consumption involving only the reactive fuel component. For instance, if there were simultaneous consumption of gasoline and diesel fuel during the initial burn, with the main heat release delay being due to increased heat transfer in the squish region for example, the fuel reaction extent would also reflect heat transfer losses, slowing the reaction progression rate in the squish region, which it does not. Finally, further confirmation of the hypothesized combustion mechanism can be seen by considering the overlaid experimental cylinder pressure and AHRR in Figure 42.

104

(B) | (B) |

(A) Fuel Reaction Extent (A) Water Reaction Extent

0.22

100

0.20

90

0.18

80

0.16

70

0.14

60

0.12

50

0.10

40

0.08

30

0.06

20

0.04

10

0.02

Pressure (bar)

110

0 -20 -16 -12

0.00 -8

-4

0

4

8

Crank Angle (deg)

12

16

20

AHRR (kJ/deg) Normalized Reaction Extent / 10 (-)

Cylinder Pressure | AHRR (B) | (A) Aldehydes Reaction Extent

Figure 42 Experimental Cylinder pressure, AHRR, and experimental optical Location A (bowl) and B (squish) aldehyde, fuel, and water reaction extent measurements.

The fuel and water reaction extents clearly show that the crank-angle at which transition into thermal ignition occurs is different depending on the local fuel reactivity. This finding provides supporting experimental evidence for the conjectured longer heat release event in the experimental work of Hanson et al. [47, 48], and the computational simulation work of Kokjohn [46]. Those studies demonstrated that dual-fuel operation could be beneficial at mid and high-loads due to the long burn durations and low pressure rise rates. To further investigate the higher load operation presented by Kokjohn and Hanson, higher load

105 operation with alternative fuels was also investigated, but without optical diagnostics, and the results are presented in Chapters 5 and 6.

106

Chapter 5: Fuel Reactivity Controlled Combustion via Reactivity enhancement 5.1

Reactivity improving chemicals

Reactivity enhancing chemicals such as 2-Ethylhexyl nitrate (2-EHN) and di-tert butyl peroxide (DTBP) have historically been used to increase fuel reactivity with diesel fuels [36-44]. These select chemicals have been found to be advantageous due to the presence of at least one low strength bond present in the molecule. The weak bond strength provides low activation energy of the initial step in the chemical mechanism. Not only are these molecules easily broken, but the initial bond cleavage is also a radical initiation step. The combination of these two chemical phenomena result in the ability of 2-EHN and DTBP to form free radicals early in the combustion process, accelerating fuel decomposition, and reducing ignition delay. Although both of these chemicals have been researched as possible ignition enhancing additives, 2-EHN has been shown to increase NOx emissions with low temperature combustion regimes. Low temperature combustion research by Ikes [40] demonstrated that if 2-EHN was used as a cetane improver, approximately 50% of the total number of fuel borne NO2 groups resulted in NOx emissions. Increases in NOx emissions are counter productive to LTC combustion regimes, which are developed as a means to reduce NOx emissions. Thus, to mitigate fuel borne NOx emissions from reactivity enhancing chemicals, it is best to avoid nitrate compounds. Peroxide compounds, like DTBP, have been shown to also offer reactivity improvement without adverse NOx penalty [39, 44]. The molecular structure of DTBP is

107 nitrogen-free, and is thus not prone to increased NOx emissions from fuel NOx. Research by Eng et al. [39] clearly demonstrates the behavior of small additions of DTBP to iso-octane in low load HCCI operation. The ignition delay trend of Figure 43 clearly shows piece-wise behavior of the ignition delay as a function of percent DTBP addition. With addition of DTBP from 0% to approximately 2% there is significant ignition delay reduction, whereas after 2%, continued addition of DTBP becomes significantly less effective. Notably the trend in ignition delay is closely mimicked by the fuel requirements for stable operation, where DTBP addition is seen to decrease fueling requirements at the light load condition. Such a trend is a resultant of the increase in fuel reactivity provided by the DTBP, where the early stages of combustion are more stable with the reactivity improver.

Figure 43 HCCI ignition delay and fueling requirement as a function of % DTBP addition to iso-octane [39].

108

The results of Eng et al. [39] show significant promise for the use of small percentage additions of DTBP as a means to increase fuel reactivity. 5.2

DTBP addition to Direct Injected Gasoline

To explore the combustion behavior trends of DTBP as described by Eng et al. [39], 2% DTBP by volume was added to gasoline with fuel properties found in Table 8 and Appendix B. Both gasoline with and without 2% DTBP were direct injected using the common rail injector described in Chapter 3. For these tests, the engine was operated at the conditions in Table 13.

Table 13 Direct Injection Operating Conditions Engine Speed (rev/min) 1300 Intake Pressure (kPa)

207.67

Exhaust Pressure (kPa)

217.80

Intake Temperature (°C)

Varied with timing

Injection Pressure (bar)

1000

Injection Timing (deg ATDC)

Varied with temperature

Maximum PRR (bar/deg)

10

Nominal IMEPn (bar)

3

Energy per Cycle (J/cycle)

1757.6

Similar to the single injection work of Hanson et al. [9], a single injection strategy was used for the tests, but the timing was varied as to retain a maximum PRR of

109 less than 10 bar/deg. Also, intake temperature was adjusted with the injection timing such that the combined effect of intake temperature and injection timing provided operation with less than 10 bar/deg. Figure 44 displays the acceptable operating injections and corresponding ignition delay (EID) for various intake temperatures. 22

E-0 E-0 + 2% DTBP

number denotes intake temperature

EID [SOIC -CA 50] (deg)

20

18 45 55 30 35 40 55 60 45 30 70 35 40 50

16

14 55

12

40

40 45

50 55

45 50

10 -14

-12

-10

-8

-6

-4

-2

0

2

4

6

Injection Timing (deg ATDC)

Figure 44 Direct injection gasoline and DTBP. Ignition delay numbers indicate intake temperature. From the figure, DTBP addition of 2 percent clearly decreases ignition delay, even requiring later injection timings to avoid adverse PRR. This provides clear supporting evidence to the HCCI work of Eng et al. [39], in which even a small percentage of DTBP added to direct injected gasoline is seen to increase the fuel reactivity significantly. Although the fuel is more reactive, the combustion duration remains relatively unchanged. Figure 45 displays the combustion trends of gasoline with and without 2% DTBP addition.

110 -16

-14

-12

-10

-8

-6

-4

-2

0

2

4

6 10 8 6

45 50 55

45 55 60 70 50

4 2 0

17 15 13 11 9 7 5

45 55 70 30

35

40 45

30 35

40

55

50 55 55

40 60

45 50

45 50

40

220 50 55 30 35

ETA thermal indicated

30 35 40

40

210

55 60 70

45 45

55

40

40 45 45 50

200

50

190

55

180 170

0.46

35

55

0.44

45

30

50

0.42

30 35

60 70

55

40

0.40 0.38 -16

45/50 40 40/45 50/55

40

ISFC gasoline eq. net (g/kW-hr)

CA50 (deg ATDC)

30 35 40

Combustion duration 10-90 (deg)

12 number denotes intake temperature E-0 Gasoline E-0 2% DTBP Gasoline 45 55 40 40 45 50 55 40 30 35

45 55

-14

-12

-10

-8

-6

-4

-2

0

2

4

6

Injection Timing (deg ATDC)

Figure 45 Combustion timing and efficiency of direct injection of gasoline at 3 (bar) IMEP operation both with and without 2% DTBP addition.

From the figure, when DTBP is added to the fuel there is a minimal impact on the combustion duration, but combustion is phased later. This is due to the inability to operate under PRR limits with early injection timings. Although the injection timings and CA 50 are later with DTBP addition, the thermal efficiency and ISFC can be slightly higher than those of gasoline without DTBP. The reasons for this are not necessarily from the combustion phasing, but also from the CO, HC, and thus combustion efficiency trends of Figure 46.

111

-8

-6

-4

-2

0

2

4

6

30 30 35 40 35 40

45

40 40 45

50 45

55 70

55

45 50

50 48

50

46

55 55

44

60

70

0.000

40 45 50 60

30

55

40 45 55 60 70 35

55 30 35 40 45 50

NOx (g/kW-hr)

30

30

8

35

40

40

45

45 50

6 35

4

40

4.0 3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0

55

40 45 50 45

10

55

45 50

50 55

70 55

40 40 45 50

60

45 50 55

55

55

2 0 -16

-14

-12

-10

-8

-6

-4

-2

0

2

4

6

Injection Timing (deg ATDC)

-6

-4

-2

0

2

4

6

1.00

40 40 45 50 55 30 30 35 40 45 50

HC (g/kW-hr)

40 40 45 50

35

-8

50

20.0 17.5 15.0 12.5 10.0 7.5 5.0 2.5 0.0

45 40

35 40

70

0.90

60

40

30

0.95

45 50 55

60 55 45 55

35 30

70

45 55 50

50 55 40 40 45 45 50 55

40 30

COV (%)

0.002

-10

35

0.006 30 30 35 40 45 55

-12

1.05

0.008

0.004

-14

number denotes intake temperature E-0 Gasoline E-0 Gasoline + 2% DTBP

42

0.010

PM (g/kW-hr)

-16 52

η comb. (-)

-10

30 35 30

45 55 40

35

700 650

40

550

35

-14

-12

-10

-8

45

35

30

500 -16

55 70

30

600

40 45

-6

20

60

40 45 55 50

50 60

-4

-2

70

55

10

40 40 45 50 45 50 55

CO (g/kW-hr)

-12

number denotes intake temperature E-0 Gasoline E-10 Gasoline + 2% DTBP

Ex. CO2 (g/kW-hr)

-14

IMEP (psi)

-16

0

50 55 40 40 45 55 45 50

0

2

4

6

Injection Timing (deg ATDC)

Figure 46 Emission and combustion stability of trends of gasoline direct injection at 3 bar IMEP operation both with and without 2% DTBP addition. The emissions and performance metrics of these two fuels can be seen in Figure 46, where 2% DTBP is seen to increase combustion efficiency, and have minimal impact on NOx and soot, and operates with low COV. The benefits of DTBP seen in the present tests with direct ignition of gasoline are thus, similar to the findings of HCCI research conducted by Eng et al. [39], where DTBP enhanced the fuel reactivity. DTBP addition to the fuel reactivity controlled combustion concept demonstrated by Kokjohn [46] and Hanson [47, 48] is explained and discussed in detail in the remainder of this section.

112 5.3

DTBP addition to achieve fuel reactivity stratification

This section presents research of reactivity stratified combustion via In-cylinder blending of port-injected pump gasoline with direct-injected pump gasoline that has been reactivity enhanced via DTBP addition. Initially, 3.5% DTBP addition by volume was added to the direct injected “reactive” fuel stream. The DTBP percentage was subsequently decreased from 3.5% to 1.75%, 0.75%, and 0% addition percents by volume. For all DTBP percentages the engine was operated with loads of 6 and 9 bar, both at an engine speed of 1300 rev/min. At the 6 (bar) condition, operation with PRR below approximately 10 bar/deg was achievable with the 3.5%, 1.75% and 0.75% DTBP addition fuels without the need for EGR. Operation with 0% DTBP addition at either the 6 or 9 (bar) loads was unrealizable, and thus no cylinder pressure or emissions data were collected. The finding that without DTBP addition at the researched conditions, combustion was unrealizable also demonstrates that the addition of DTBP successfully established the reactivity gradient that is required for fuel reactivity controlled combustion. As shown in the previous section, the gradient enables combustion progression from areas of higher to lower local fuel reactivity, thus extending the heat release event. The inability for operation with only a single fuel of low reactivity demonstrates that equivalence ratio effects are minimal for this combustion regime. 5.3.1

Low Load DTBP Study

Initially the effect of DTBP addition was studied at the light load condition of 6 (bar) IMEP. The light load condition was selected as the starting point because

113 EGR was not required for low PRR operation. Operation without EGR enables investigation of only the fuel reactivity effects without confounding effects due to any potential interactions between EGR and DTBP, or EGR diluent effects that could result in unknown engine performance. 5.3.2

DTBP percent sweep with 0% EGR

The 6 (bar) IMEP operating conditions can be seen in Table 14.

114

Table 14 6 (bar) IMEP baseline operating conditions Total fueling (kg/hr) 2.72 EGR (%)

0

Intake Temperature (°C)

40

Port fueling percentage (%)

90

Port injection pressure (bar)

4.14

Direct injection pressure (bar)

400

SOI1 fuel percentage (%)

60%

SOI2 fuel percentage (%)

40%

Port Injection Timing (° ATDC)

-320

SOI1 (first direct injection) Command -55 Timing (° ATDC) SOI2 (second direct injection) Command -36 Timing (° ATDC)

Of

particular

Intake Pressure (kPa)

174.4

Exhaust Pressure (kPa)

184.1

note,

two of

the

operating parameters

of

the

present

gasoline/gasoline with DTBP direct-injection tests are different than those of the aforementioned gasoline/diesel 4.5 (bar) optical tests. For operation with gasoline/gasoline DTBP, the injection pressure was reduced to 400 (bar) from 800 bar. Preliminary operation at both 800 (bar) and 600 (bar) demonstrated that injection timings were unable to be fully advanced, thus non-optimally preparing

115 the charge or reactant gradient. For example, the 600 (bar) injection cases presented in Figure 47 have a SOI2 (second injection) timing of -22° ATDC. However, at the lower injection pressure of 400 bar, SOI2 injection can be optimally timed at -36° ATDC, similar to the condit ions used by Kokjohn [46] and Hanson [47,48], and those of the previously presented optical tests in Chapter 3. The reason for this injection timing limitation is likely a result of fuel distillation differences between gasoline and diesel fuels. Shi [71] has computationally shown that the low boiling point of gasoline results in rapid evaporation when direct injected. Once evaporated, the injection penetration is diminished. This results in under mixing of the injected fuel, thereby changing the reactivity gradient within the combustion chamber. As the injection event is advanced further, the penetration and mixing process remain unchanged, causing similar under-mixed behavior, but the advanced injection timing allows additional fuel temperature history. The additional temperature history, combined with the locally under-mixed richer than desirable region, promotes ignition advance, and thus higher PRR result. Although this was experimentally observed, supporting data for these higher than desirable PRR operation case was not captured, due to engine mechanical and instrumentation considerations. Furthermore, the under-mixed fuel rich areas promote higher local temperatures and NOx emissions. If the injection pressure is reduced, the droplet size of the direct injected fuel increases [71], and thus slows the droplet evaporation process through the smaller surface-to-volume ratio of the spray liquid (~3/r). The lengthened

116 evaporation time enables the injection to both penetrate further, and more fully mix and distribute throughout the combustion chamber. The result of this effect can be seen in Figure 47 where operation at both 600 (bar) and 400 (bar) injection pressures is displayed. Note that the tests in Figure 47 display slightly higher PRR in the 400 (bar) injection pressure case. The increased PRR with 400 (bar) injection pressure at 40°C is thought to be an effect of the lower IMEP between the two cases. The 600 (bar) injection pressure was operated slightly below 6 (bar) IMEP, and aided in lowering PRR.

3.5% DTBP Gasoline (600 (bar) Injection Pressure) 3.5% DTBP Gasoline (400 (bar) Injection Pressure)

110 100

1.00

90 80

0.95

7.5 5.0 2.5 0.0

450

40

41

42

43

44

45

46

47

48

Intake Temperature (°C)

49

50

51

6 4 4.0 3.0 2.5 2.0 1.5

0

500

8

3.5

NOx (g/kW-hr)

Ex. CO2 (g/kW-hr)

550

10

2

20

10

400 39

70

PRR (bar/deg)

0.90

CO (g/kW-hr)

HC (g/kW-hr)

10.0

IMEP (psi)

η comb. (-)

1.05

COV (%)

3.5% DTBP Gasoline (600 (bar) Injection Pressure) 3.5% DTBP Gasoline (400 (bar) Injection Pressure)

1.0

1.2 1.0 0.8 0.6 0.4 0.2 0.0 39

2010 limit

40

41

42

43

44

45

46

47

48

49

50

51

Intake Temperature (°C)

Figure 47 Effect of direct-injected gasoline injection pressure and intake temperature. Note the significant difference in NOx emissions with similar combustion behavior between 600 and 400 (bar) injection pressures. The second parameter that differed from the low load optical tests in the previous section, was the intake temperature. For the 6 (bar) IMEP condition, the intake

117 temperature was raised to 40°C from 32°C used in th e optical tests. Unlike the optical tests where no EGR and low PRR were desired, it was desired with these tests to investigate the effect of EGR with DTBP. Preliminary testing found that 40°C intake temperature allowed good combustion sta bility with the addition of EGR. The effect of EGR at the 6 (bar) condition will be discussed in section 5.1.6 to follow. The pressure and AHRR traces from operation at 6 (bar) IMEP with the DTBP percentages of 3.5%, 1.75% and 0.75%, can be seen Figure 48.

0% EGR 40°C Intake Temperature

120

1.2

0.05 0.04 0.03 0.02 0.01 0.00 -25

1.0

-20 -15 -10 -5 Crank Angle (deg)

0.8

0

3.5 1.75 0.75

60

0.6

40

0.4 % DTBP

20

0.2

0

0.0

-25 -20 -15 -10

-5

0

5

10

15

20

AHRR (kJ/deg)

Pressure (bar)

80

AHRR (kJ/deg)

NTC behavior

100

25

Crank Angle (deg) Figure 48 Pressure and AHRR traces for in-cylinder blended fuel reactivity stratification via DTBP addition to pump gasoline at 6 (bar) load. From Figure 48 it can be seen that decreasing the DTBP percentage results in slightly later combustion phasing. The slight difference in the magnitude of the

118 heat release peak of the 1.75% case is evident because the fueling was slightly lower than the other cases, as seen in Table 15, but all other parameters were identical to those of Table 14. Table 15 6 (bar) IMEP DTBP percentages fueling DTBP (%) 0.75 1.75 3.5 Total Fueling (kg/hr)

2.70

2.66

2.72

Percent Port Fuel (%)

90

90

90

The slightly lower fueling was inadvertently caused by slightly lower port fueling pressure. This slight decrease in fuel energy can be seen in the lower magnitude of the heat release rate peak of the 1.75% case, and slightly lower expansion pressure. However, the combustion timings and trends of the DTBP sweep are consistent with those of Eng et al. [39]. The trend is consistent with their observations that decreases to the DTBP percentage beyond approximately 2% results in significant combustion behavior differences. Figure 49 provides further insight to the pressure and heat release rate trends of Figure 48 above.

119

0.5

1.0

1.5

2.0

2.5

3.0

3.5

4.0 11

Combustion duration 10-90 (deg)

0.0

8

9

6 4 2 0 -2

170 165

ETA thermal indicated

160 155 150 0.54

ISFC net (g/kW-hr)

CA50 (deg ATDC)

10

0.52 0.50 0.48 0.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

4.0

Percent DTBP

Figure 49 Combustion behavior trends of various DTBP percentage additions to gasoline, for the tests in Table 15.

5.3.3

EGR Sweep

In addition to fuel reactivity, EGR effects were investigated at the 6 (bar) IMEP condition. EGR effects were only investigated with the most reactive fuel condition of 3.5% DTBP addition. The most reactive case was chosen because out of all of the tested DTBP percentages, it was assumed that the increased reactivity would be the most tolerant to EGR, and it was unknown how EGR tolerant DTBP operation would be. The operating conditions for the EGR sweep

120 were identical to those for the DTBP percentage sweep previously described in Table 14. The Cylinder pressure and AHRR traces from the sweep can be seen in Figure 50 below, where increasing EGR results in later combustion phasing, and lengthened heat release event. The 3.5% DTBP addition was shown to be very tolerant to EGR allowing up to 35% EGR, even at a moderately light load.

40°C Intake Temperature EGR Sweep

120

1.2

0.05 0.04 0.03 0.02 0.01 0.00 -25 -20 -15 -10 -5 Crank Angle (deg)

1.0 0.8

0

0% 8%

60

0.6

14% 20% 25% 30%

40

0.4

35%

20

AHRR (kJ/deg)

Pressure (bar)

80

AHRR (kJ/deg)

NTC behavior

100

0.2 % EGR 0.0

0

-25 -20 -15 -10

-5

0

5

10

15

20

25

Crank Angle (deg) Figure 50 EGR sweep with 3.5% DTBP at 6 (bar) net IMEP The combustion and emissions trends of both the DTBP and EGR sweeps are simultaneously plotted in Figure 51, where both the beneficial effects of fuel reactivity gradients and EGR on combustion phasing and emissions are clearly seen.

121

HC (g/kW-hr)

10 8 6 4 2

2.5 0.3

2.0

0.2

2010 limit

0.1 0.0 0

5

10

15

20

EGR(%)

25

30

35

Ex. CO2 (g/kW-hr)

3.0

0.92

7.5 5.0 2.5

4.0 3.5

NOx (g/kW-hr)

10.0

0.000

17 15 13 11 9 7 5

500 480 460 440 0

5

10

15

20

EGR(%)

25

30

35

Combustion duration 10-90 (deg)

0.94

13 11 10

8

9

6 4 2 0 -2

170 165 160 155 150

0.54

ISFC net (g/kW-hr)

0.005

14 12

CA50 (deg ATDC)

0.96

COV (%)

PRR (bar/deg)

12

0.98

0.010

15

gasoline/3.5%DTBP gasoline gasoline/1.75%DTBPgasoline gasoline/0.75%DTBPgasoline

η comb. (-)

PM (g/kW-hr)

0.015

2010 limit

1.00

gasoline/3.5%DTBP gasoline gasoline/1.75%DTBP gasoline gasoline/0.75%DTBP gasoline

ETA thermal indicated

0.020

CO (g/kW-hr)

gasoline/3.5%DTBP gasoline gasoline/1.75%DTBP gasoline gasoline/0.75%DTBP gasoline

0.52 0.50 0.48 0

5

10

15

20

25

30

35

EGR(%)

Figure 51 Combustion behavior and emissions trends with EGR addition and fuel reactivity differences. Note the broad high efficiency, low pressure rise rate operation potential with DTBP addition and EGR.

122 From Figure 51, as EGR is increased combustion phases later, and NOx, CO, HC, PRR, and PM simultaneously decrease. Concurrently, ISFC, thermal efficiency, COV, and exhaust CO2 remain relatively undisturbed, and only deviate once the highest level of EGR was tested, at which the onset of combustion instability was observed by the turn-up of COV to approximately 4%. 5.3.4

9 (bar) operation

Building on the 6 (bar) IMEP results, the load was increased to 9 (bar) IMEP. Unlike the 6 (bar) condition, operation at 9 (bar) required the addition of EGR to reduce pressure rise rates below approximately 10 bar/deg. The operational conditions at the 9 (bar) condition were matched to those of the 9 (bar) IMEP high efficiency operation demonstrated by Hanson [47, 48] in Table 16.

123

Table 16 9 (bar) net IMEP baseline operating conditions Total fuel energy (J/cycle) 4044.7 Port fueling percentage (%)

89 up to 94

EGR (%)

43

Intake Temperature (°C)

32

Port injection pressure (bar)

4.14

Direct injection pressure (bar)

400

SOI1 fuel percentage (%)

60%

SOI2 fuel percentage (%)

40%

Port Injection Timing (° ATDC)

-320

SOI1 Command Timing (° ATDC)

-55

SOI2 Command Timing (° ATDC)

-36

Intake temperature (°C)

38

Intake Pressure (kPa)

174.4

Exhaust Pressure (kPa)

184.1

The comparison between gasoline/gasoline with DTBP and gasoline/diesel was made to investigate the behavioral differences of DTBP addition to gasoline in compression ignition. Similar to the 6 (bar) IMEP tests, the DTBP percentage was swept in levels of 3.5%, 1.75%, 0.75%. As previously discussed, attempts to operate the 6 (bar) condition with 0% DTBP were unsuccessful, and thus were not attempted at the higher load of 9 bar. For all DTBP percentage tests, 89% port fuel of gasoline was the initial test condition. Eighty nine percent was

124 selected for two reasons. Primarily, it was found to have the best thermal efficiency from Hanson [47, 48]. Secondly, it was desired to hold the PRR under 10 bar/deg, and when the port fueling was reduced below 89%, PRR increased accordingly. Except for the port fueling percentage, and unless otherwise noted, all the 9 (bar) operating condition parameters were held constant. The port fueling percentage was swept from 89% to as much as 94%, to study the sensitivity to and the effect of the DTBP additive-induced reactivity gradient. While sweeping port fueling percentages, the direct injection durations of SOI1 and SOI2 were reduced in proportion to most accurately retain the relative distribution of the reactivity gradient for all DTBP percentages. 5.3.5

3.5% DTBP addition

The first DTBP percentage tested at the 9 (bar) IMEP condition was the most reactive addition case of 3.5% DTBP to the direct-injected gasoline fuel. The percentages of port fuel tested at this condition were 89%, 91%, and 92%, respectively. The pressure and AHRR of this port fueling sweep with 3.5% DTBP are seen in Figure 52, with the combustion performance and property trends shown in Figure 53. For both Figures 52 and 53 the results of Hanson [47, 48] are plotted in black for comparison.

125

3.5% DTBP

140

1.4

100

1.2

0.05 0.04 0.03

1.0

0.02 0.01 0.00 -20

-15 -10 -5 Crank Angle (deg)

80

0

0.8

89 91

60

0.6

92

Gasoline Diesel 89%

40

0.4

baseline*

% port fuel

20

0.2

0

0.0

-25 -20 -15 -10

-5

0

5

10

15

20

AHRR (kJ/deg)

Pressure (bar)

120

AHRR (kJ/deg)

NTC behavior

25

Crank Angle (deg) * data from R.M.Hanson

Figure 52 Indicated pressure and AHRR trends of 3.5% DTBP addition to direct injected gasoline in fuel reactivity controlled combustion at 9 (bar) IMEP, compared to the highest efficiency case of Hanson [47, 48], (black). Combustion duration 10-90 (deg)

12

gasoline/3.5 % DTBP gasoline gasoline/diesel (R.M. Hanson)

11 10 9 8 6

8 6 4 2 0 -2

170 165

ETA thermal indicated

160 155 150 0.54

ISFC net (g/kW-hr)

CA50 (deg ATDC)

7 10

0.52 0.50 0.48 0.81

0.83

0.85

0.87

0.89

0.91

0.93

0.95

percent port fuel (%)

Figure 53 Combustion trend of 3.5% DTBP addition to direct injected gasoline in fuel reactivity controlled combustion at 9 (bar) IMEP, as compared to cases of Hanson [47, 48], (black).

126

From the results presented in Figures 52 and 53 it is clear that DTBP addition of 3.5% to the direct-injected gasoline exhibits similar main heat release behavior to that of direct-injected diesel when operating with reactivity controlled combustion. The combustion timings of the 89% port fuel percentage with the 3.5% DTBP case is slightly before, and the 91% case is slightly after the gasoline/diesel case of Hanson. Although there was not a 90% DTBP case, it is assumed that it would reside between the 89% and 91% pressure and AHRR trances. This would then be best representative of the combustion timings of the gasoline/diesel traces of Hanson. From Figure 52 the combustion duration of the DTBP addition is slightly shorter than that of the gasoline/diesel cases of Hanson, but the increasing combustion duration with port fuel percentage trend is similar. The slightly shorter combustion duration would be a resultant of differences in the combustion chemistry species and thus reaction mechanisms. Chemical differences are particularly evident in the early stages of the heat release event, which are critical to compression ignition combustion. When examining the early stages of the heat release in Figure 52 there is a

notable

difference

between

the

behavior

of

gasoline/diesel

and

gasoline/gasoline with DTBP. The DTBP additized gasoline exhibits early heat release that occurs approximately 3 crank angles before that of the diesel case, and is approximately half of the peak magnitude observed with diesel. This low magnitude early negative temperature coefficient (NTC) behavior with DTBP addition is similar to the results demonstrated by Eng et al. [39]. That research demonstrated that while operating with traditional HCCI, DTBP negative

127 temperature coefficient NTC behavior occurred earlier and with lower magnitude than n-heptane. The authors of that study demonstrate that this different behavior results in different chemical paths than n-heptane. Unlike diesel/n-heptane NTC where not only radicals are formed but heat is also released [29], the early cycle heat release with DTBP is primarily responsible for only forming radicals and not simultaneously releasing significant heat. The observations of Eng et al. [39], seen with traditional HCCI operation were also seen in the present fuel reactivity-controlled combustion operation. The pressure trace trends in Figure 52 indicate that the cylinder pressure during the early NTC stages of combustion is lower with DTBP than that of the gasoline/diesel case. Because the operating conditions are identical between the two cases the lower compression pressure with DTBP is directly a result of the lower bulk cylinder temperature. This lower cylinder pressure prior to TDC is also reflected in the thermal efficiency trends in Figure 53. The slight decrease in negative compression work before TDC with DTBP increases net thermal efficiency and lowers ISFC as compared to diesel. Table 17 numerically displays the indicated gross work, power, and thermal efficiency of the gasoline/gasoline 3.5% DTBP, and is compared to the gasoline/diesel operation of Hanson [47, 48].

128

Table 17 Gross cycle performance comparison of Diesel and 3.5% DTBP 9 (bar) operation with matched conditions Direct Injection Reactivity Fuel Diesel 3.5% DTBP Port Fuel Percentage (%) 89 89 -8.83 -8.65 Indicated Compression pressure (bar) Indicated Expansion pressure (bar) Indicated Gross Power (kW) Indicated Gross Thermal Efficiency (%)

18.02

18.01

24.28

24.74

0.562

0.572

In comparing the emissions between operation with gasoline/gasoline with DTBP and gasoline/diesel, there is good agreement in the overall emissions trends. The indicated emissions trends can be seen in Figure 54 for both fuels. The DTBP appears to generally produce lower PM, and CO emissions with comparable COV of IMEP, PRR, but with slightly higher NOx, and HC emissions. It is believed that these trends are not from the addition of DTBP, but from operation with direct-injection of gasoline and not diesel. (Further evidence to support this claim is seen later by the near identical emissions trends between all tested DTBP percentages in Figure 61.)

129

0.020

0.98 0.97 5

0.000

26 22 18 14 10 6 2

HC (g/kW-hr)

PRR (bar/deg)

0.005

0.99

η comb. (-)

0.010

4.0

0.96

4 3 2 1 0

7

3.5 2.5 2.0

6

COV (%)

3.0

5 4 3

NOx (g/kW-hr)

0.3

1.0

0.2

2010 limit 0.1 0.0 0.81

0.83

0.85

0.87

0.89

0.91

0.93

0.95

Ex. CO2 (g/kW-hr)

1.5 550

2

540

1

CO (g/kW-hr)

0.015

2010 limit

1.00

gasoline/3.5% DTBP gasoline gasoline/diesel (R.M. Hanson)

PM (g/kW-hr)

gasoline/3.5 % DTBP gasoline gasoline/diesel (R.M. Hanson)

530 520 510 500 0.81

percent port fuel (%)

0.83

0.85

0.87

0.89

0.91

0.93

0.95

percent port fuel (%)

Figure 54 Engine-out emissions of fuel reactivity-controlled combustion with gasoline/gasoline with DTBP and those of gasoline/diesel from Hanson [47, 48]. Interestingly in Figure 54, the combustion efficiency of operation with gasoline/gasoline

with

DTBP

is

slightly

diminished

as

compared

to

gasoline/diesel, a result of increased HC emissions. It is unknown what the particular reasons for these increases are, but they could be due to either one or the combination of un-reacted crevice flows and end of injection dribble present with the lower injection pressure in the direct injection events. If these combustion losses could be reduced, for example by operating with increased coolant temperature, the thermal efficiency of DTBP operation would increase further. Although the combustion behavior of operation with gasoline and gasoline with 3.5% DTBP addition was close to that of gasoline/diesel, the combustion duration was decreased. As the DTBP percentage is reduced it

130 would expected that combustion would approach that of a single fuel PRR and combustion properties. Thus, this was explored next. 5.3.6

1.75% DTBP addition

After successfully demonstrating operation with 3.5% DTBP addition to direct injected gasoline, the percent of DTBP was reduced to 1.75% DTBP. With reference to Figure 43 reproduced from Eng et. al. [39], 1.75% DTBP is approximately at the intersection of the piece-wise behavior observed in their research. The experimental fuel reactivity controlled cylinder pressure and AHRR traces from operation with 1.75% DTBP can be seen in Figure 55, where the baseline case of Hanson [47, 48] is again plotted in black.

1.75% DTBP

140

1.4

NTC behavior

1.2

0.04 0.03 0.02

1.0

0.01 0.00 -20

-15 -10 -5 Crank Angle (deg)

0

89 90

80

0.8 92 93 94

60

0.6

Gasoline Diesel 89%

40

baseline*

% port fuel

0.4

20

0.2

0

0.0

-25 -20 -15 -10

-5

0

5

10

Crank Angle (deg)

15

20

AHRR (kJ/deg)

Pressure (bar)

100

AHRR (kJ/deg)

0.05

120

25

* data from R.M.Hanson

Figure 55 Indicated pressure and AHRR trends of 1.75% DTBP addition to direct-injected gasoline in fuel reactivity-controlled combustion at 9 (bar) IMEP, compared to the highest efficiency case of Hanson [47, 48], (black).

131 From Figure 55, operation with 89% port fuel and 1.75% DTBP results in combustion prior to that of the gasoline/diesel results of Hanson [47, 48]. This is similar to the previously discussed trend with 3.5% DTBP. However, with 1.75% DTBP a 90% port fueling condition was tested. As postulated in the analysis of the 3.5% DTBP case, the 90% port fueling case nearly matches the behavior of fuel reactivity controlled combustion with gasoline/diesel. Similar to the results of 3.5% DTBP, combustion can easily be phased by adjusting the port fueling percentage, and thus altering the reactivity gradient. With 1.75% DTBP, successful engine operation up to 94% port fuel was demonstrated. To directly compare the sensitivity of operation with 3.5% and 1.75% DTBP additions, the pressure and AHRR of identical port fueling conditions of the two tests were overlaid, Figure 56 displays the results of the comparison.

132 140

1.4

100 80

1.2

0.05 0.04 0.03

1.0

0.02 0.01 0.00 -20

-15 -10 -5 Crank Angle (deg)

0

0.8

89

89% baseline* 91

60

0.6

92

40

0.4 % port fuel

Dashed Lines 1.75% DTBP

20

Solid Lines 3.5 % DTBP

AHRR (kJ/deg)

Pressure (bar)

120

AHRR (kJ/deg)

NTC behavior

0.2

0

0.0

-25 -20 -15 -10

-5

0

5

10

15

20

25

Crank Angle (deg) * data from R.M.Hanson

Figure 56 Comparison of 3.5% (solid colored lines) and 1.75% (dashed colored lines) DTBP with identical port-fueling percentages, and compared to the highest efficiency case of Hanson [47, 48], (black).

As seen from Figure 56, the heat release events of the respective port fueling percentages between 3.5% and 1.75% DTBP addition are nearly identical in timing and duration. However, the 1.75% DTBP exhibits peak pressure and AHRR that are slightly higher in magnitude that that of 3.5% DTBP. This increase in peak magnitude would be expected because as the DTBP percentage is decreased, the combustion approaches that of a single fuel reactivity, and would be consistent with the combustion behavior observed in conventional, externallyblended single reactivity fueled HCCI. Supporting this observation are the combustion trends of Figure 57, where the two gasoline with DTBP percentages

133 and gasoline/diesel cases are depicted. From the figure, all of the combustion trends are nearly identical between the 3.5% and 1.75% DTBP cases, except for the 10-90% combustion duration, which decreased as the percentage of DTBP is decreased. This evidence along with the higher peak AHRR magnitudes observed in Figure 57 supports the trend that as the DTBP percentage is reduced, the combustion behaves more similar to that of traditional HCCI a with single externally blended fuel reactivity. 11 10 9 8 7 10

6

8 6 4 2 0 -2

170 165

ETA thermal indicated

160 155 150 0.54

ISFC net (g/kW-hr)

CA50 (deg ATDC)

Combustion duration 10-90 (deg)

12

gasoline/3.5 % DTBP gasoline gasoline/3.5 % DTBP gasoline gasoline/diesel (R.M. Hanson)

0.52 0.50 0.48 0.81

0.83

0.85

0.87

0.89

0.91

0.93

0.95

percent port fuel (%)

Figure 57 Combustion trend of 3.5% and 1.75% DTBP addition to direct injected gasoline in fuel reactivity-controlled combustion at 9 (bar) IMEP, compared to cases of Hanson [47, 48].

With the similar combustion timing and behavior shown in Figure 57, the emissions of the 1.75% DTBP would be expected to be nearly identical to those of the 3.5% DTBP cases. Figure 58 demonstrates this relation, where all of the emissions trends are very similar, except for CO, and thus the combustion

134 efficiency. The CO emissions have a hooked behavior, where CO emissions start high, decrease, and then again increase. Although this trend is interesting, the overall magnitude of the trend is very similar to the other fuels tested, and is not thought to be significant evidence for major differences in combustion behavior between the fuels. The observed COV of the 1.75% case is slightly lower than that of the 3.5% DTBP case, but the values are similar to those of gasoline/diesel operation seen by Hanson [47, 48]. A sharp upturn in COV was observed, where reduced combustion stability was observed once the port fueling percentage was increased past 92%.

0.99

η comb. (-)

0.005

0.98 0.97

0.000

5

HC (g/kW-hr)

26 22 18 14 10 6 2

1.00

4.0

0.96

4 3 2 1 7

0

3.5 2.5 2.0

6

COV (%)

3.0

5 4 3

NOx (g/kW-hr)

0.3

1.0

0.2

2010 limit 0.1 0.0 0.81

0.83

0.85

0.87

0.89

0.91

percent port fuel (%)

0.93

0.95

Ex. CO2 (g/kw-hr)

1.5 550

2

540

1

CO (g/kW-hr)

0.010

PM (g/kw-hr)

0.015

2010 limit

PRR (bar/deg)

gasoline/3.5% DTBP gasoline gasoline/1.75% DTBP gasoline gasoline/diesel (R.M. Hanson)

0.020

gasoline/3.5% DTBP gasoline gasoline/1.75% DTBP gasoline gasoline/diesel (R.M. Hanson)

530 520 510 500 0.81

0.83

0.85

0.87

0.89

0.91

0.93

0.95

percent port fuel (%)

Figure 58 Engine-out emissions of fuel reactivity controlled combustion with gasoline/gasoline with DTBP percentage of 3.5% and 1.75%, and those of gasoline/diesel from Hanson [47, 48].

135 From the results of 1.75% and 3.5% DTBP addition to pump gasoline, there appears to be good agreement between DTBP addition percentages above 1.75%. The results showed no significant combustion or emissions differences between

both

DTBP

percentage,

and

only

slight

disagreement

from

gasoline/diesel results of Hanson [47, 48]. To further explore the effect of DTBP addition, the DTBP percentage was further decreased to explore the reduced ignition delay behavior seen by Eng et al. [39]. 5.3.7

0.75% DTBP addition

The DTBP percentage was decreased from 1.75% to 0.75% DTBP addition. At this percentage, significant deviation from the operational behavior of the 3.5% and 1.75% addition cases was observed. As described by Eng et al. [39], with DTBP percentages less than approximately 2%, the reactivity enhancing effect of DTBP is dramatically reduced. The indicated pressure and AHRR traces in Figure 59, and the combustion properties in Figure 60 also clearly demonstrate this behavior.

136 0.75% DTBP

140

1.4

NTC behavior

1.2

0.04 0.03 0.02

1.0

0.01 0.00 -20

-15 -10 -5 Crank Angle (deg)

0

89%

80

0.8

90% 92%

60

0.6 94%

Gasoline Diesel 89%

40

baseline*

% port fuel

0.4

20

0.2

0

0.0

-25 -20 -15 -10

-5

0

5

10

15

Crank Angle (deg)

20

AHRR (kJ/deg)

Pressure (bar)

100

AHRR (kJ/deg)

0.05

120

25

* data from R.M.Hanson

Figure 59 Indicated pressure and AHRR trends of 0.75% DTBP addition to direct-injected gasoline in fuel reactivity-controlled combustion at 9 (bar) IMEP, compared to the highest efficiency case of Hanson [47, 48], (black). gasoline/3.5 % DTBP gasoline (43% EGR) gasoline/1.75 % DTBP gasoline (43% EGR) gasoline/0.75 % DTBP gasoline (43% EGR) gasoline/diesel (R.M. Hanson)

11 10 9 8 7

10

6

8 6 4 2 0 -2

170 165

ETA thermal indicated

160 155 150 0.54

ISFC net (g/kW-hr)

CA50 (deg ATDC)

Combustion duration 10-90 (deg)

12

0.52 0.50 0.48 0.81

0.83

0.85

0.87

0.89

0.91

0.93

0.95

percent port fuel (%)

Figure 60 Combustion trends of 3.5%, 1.75%, and 0.75% DTBP addition to direct-injected gasoline in fuel reactivity-controlled combustion at 9 (bar) IMEP, compared to cases of Hanson [47, 48].

137 As seen in Figure 60 as the DTBP percentage is decreased from 1.75% to 0.75% the combustion duration continues to decrease, as also observed with the decrease from 3.5% to 1.75% DTBP. However, unlike the decrease of DTBP percentage from 3.5% to 1.75%, decreasing the DTBP percentage from 1.75% to 0.75% resulted in adverse impacts on the combustion CA 50, ISFC, and thermal efficiency. Decreasing the DTBP percentage from 1.75% to 0.75% phased combustion slightly later; as observed by the offset in CA 50 timing, diminished thermal efficiency, and increased ISFC. The jog in the thermal efficiency and in ISFC trends at the 89% port fuel condition are further explained by the COV and PRR trends in Figure 61.

0.005

0.99 0.98 0.97

0.000

5

HC (g/kW-hr)

26 22 18 14 10 6 2

1.00

η comb. (-)

0.010

PM (g/kw-hr)

0.015

2010 limit

PRR (bar/deg)

gasoline/3.5% DTBP gasoline gasoline/1.75% DTBP gasoline gasoline/0.75% DTBP gasoline gasoline/diesel (R.M. Hanson)

0.020

4.0

0.96

4 3 2 1 7

0

3.5 2.5 2.0

6

COV (%)

3.0

5 4 3

NOx (g/kW-hr)

0.3

1.0

0.2

2010 limit

0.1 0.0 0.81

0.83

0.85

0.87

0.89

0.91

percent port fuel (%)

0.93

0.95

Ex. CO2 (g/kW-hr)

1.5 550

2

540

1

CO (g/kW-hr)

gasoline/3.5% DTBP gasoline gasoline/1.75% DTBP gasoline gasoline/0.75% DTBP gasoline gasoline/diesel (R.M. Hanson)

530 520 510 500 0.81

0.83

0.85

0.87

0.89

0.91

0.93

0.95

percent port fuel (%)

Figure 61 Engine-out emissions of fuel reactivity-controlled combustion with gasoline/gasoline with DTBP 3.5%, 1.75%, and 0.75%, compared to those of gasoline/diesel from Hanson [46, 47].

138 The higher COV and PRR observed with low port fuel percentages of 0.75% DTBP in Figure 61, demonstrates that the combustion behavior is becoming less stable and unpredictable. As each cycle drifts, combustion timing becomes less than optimal. The result of the non-optimal combustion timings is diminished power and thermal efficiency, and thus increased fuel consumption. Also, the lower indicated power of the 89% port fuel case not only confirms this, but also results in higher emissions trends from lower indicated power in the denominator of the indicated emissions calculations. Clearly from the PRR trends, operation with 0.75% DTBP is not practical unless higher port fuel percentages are used to phase combustion later. As a result, combustion efficiency suffers, as does thermal efficiency through reduced combustion efficiency, increased volume for heat transfer, and decreased expansion volume for work extraction. In addition to the combustion behavior trends, the AHRR traces in Figure 59 display increased late cycle anomalies, such as ringing in the AHRR. From these trends it is seen that decreasing the DTBP addition past approximately 1.75% results in dramatic combustion behavior differences, which are less desirable. This trend was verified through the aforementioned attempt to operate the engine with 0% DTBP addition. With 0% DTBP addition, successful operation with the conditions in Table 14 at 6 (bar) IMEP was unrealizable. Modification to the injection timings durations, and fuel percentage splits were attempted as a means to establish reasonable combustion stability with 0% DTBP addition, but were ultimately unsuccessful. From the inability to operate at 6 (bar) IMEP with 0% DTBP, load increase to 9 (bar) IMEP was not attempted.

139 5.3.8

89% and 90% port fuel combustion comparisons

The preceding sections have shown that fuel reactivity-controlled combustion using direct injection with DTBP addition to a single fuel stock of gasoline is an effective alternative to the dual gasoline and diesel fuel demonstrated in Chapter 3 and by Hanson [47, 48]. However, as discussed above with identical port fueling amounts of 89% and port gasoline; the main heat release of gasoline/diesel reactivity controlled combustion is phased slightly later, with lower peak heat release magnitude. Figure 62 displays the trend by re-plotting the indicated pressure and AHRR traces of 89% operation.

89% Port Fuel 140

1.4 NTC behavior

1.2

0.04 0.03 0.02

1.0

0.01 0.00 -20

-15 -10 -5 Crank Angle (deg)

0

0.75

80

0.8

1.75 3.5

60

0.6 Gasoline Diesel

40

0.4

baseline*

% DTBP

20 0

AHRR (kJ/deg)

Pressure (bar)

100

AHRR (kJ/deg)

0.05

120

0.2 0.0

-25 -20 -15 -10

-5

0

5

10

Crank Angle (deg)

15

20

25

* data from R.M.Hanson

Figure 62 Identical port fueling percentages between gasoline with DTBP and diesel of [47, 48] (black), at 9 (bar) IMEP.

140 The slight advance in combustion timing suggests that once initiated, DTBP addition to gasoline transitions into main heat release more rapidly than gasoline/diesel. This suggests that although DTBP appears to behave similarly to diesel in the main heat release phase, there is a difference in the chemical path in the transition into main heat release between DTBP and diesel. This reflects differences in NTC behavior of diesel as compared to the diminished and advanced NTC behavior of gasoline with DTBP. Secondly, the decreased magnitude of the main heat release with the baseline gasoline/diesel operation is not from a difference in fueling between the cases, but from both the increased early cycle consumption of fuel during NTC and the lower 10-90% time of diesel as compared to DTBP. As previously mentioned, the difference in NTC behavior of the two reactivity-enhancing fuels is significant. As seen from Figure 63, when the port fueling percentage is increased to 90% the pressure and AHRR traces of gasoline/diesel and gasoline/gasoline with DTBP display very similar main heat release timings. Although data at the 90% port fueling percentage with 3.5% DTBP was not taken, the trends observed at 89% suggest that as the DTBP percentage is increased, the main heat release magnitude would decrease, and with sufficient DTBP addition would approach that of the diesel baseline.

141

90% Port Fuel 140

1.4

NTC behavior

1.2

0.04 0.03 0.02

1.0

0.01 0.00 -20

-15 -10 -5 Crank Angle (deg)

0

0.75

80

0.8

1.75

60

0.6 89% baseline*

40

% DTBP

0.4

20

0.2

0

0.0

-25 -20 -15 -10

-5

0

5

10

15

Crank Angle (deg)

20

AHRR (kJ/deg)

Pressure (bar)

100

AHRR (kJ/deg)

0.05

120

25

* data from R.M.Hanson

Figure 63 Similar positions of main heat release between gasoline/gasoline with DTBP and gasoline/diesel [47, 48] (black), shown at 9 (bar) IMEP. Additionally from Figure 63, further evidence for later combustion phasing with the closer to single fuel reactivity behavior at the lower DTBP addition level of 0.75% is observed. If operation with less DTBP had been tested, it is reasonable to expect that, combustion would be phased later yet and a higher maximum heat release rate would be seen. 5.3.9

Operation with DTBP and Increased EGR level

To reduce PRR and to examine the EGR tolerance of DTBP operation at 9 (bar) IMEP operation with 50% EGR was also explored. To counteract the retarded combustion phasing effects of EGR, at the 50% EGR condition the intake temperature was increased to 38°C from 32°C. Althou gh intake temperature was

142 used, there is no reason to expect that 32°C operat ion at 50% EGR could not be appropriately phased by decreasing the port fuel percentage, thus advancing combustion through increased fuel reactivity instead of increased intake temperature. Operation at 50% EGR was only proven at the most reactive tested DTBP addition percentage of 3.5%. The combustion behavior of this sweep is demonstrated in Figure 64, where identical CA 50 timings are plotted for both operation at 50% and 43% EGR with 3.5% DTBP and the gasoline/diesel baseline of Hanson [47, 48]. Figure 65 displays the combustion performance trends of the baseline gasoline/diesel case of Hanson with 43% and 50% EGR and 3.5% DTBP.

140

3.5% DTBP with 43% and 50% EGR

1.4

80

1.2

0.05 0.04 0.03

1.0

0.02 0.01 0.00 -20

-15 -10 -5 Crank Angle (deg)

0

0.8

89 solid / 77 dot 91 solid / 81 dot

60

92 solid / 83 dot

0.6

Gasoline Diesel 89%

40

0.4

baseline*

% port fuel

20

Solid lines 43% EGR Dotted lines 50% EGR

AHRR (kJ/deg)

Pressure (bar)

100

AHRR (kJ/deg)

NTC behavior

120

0.2

0

0.0

-25 -20 -15 -10

-5

0

5

10

15

20

25

Crank Angle (deg) * data from R.M.Hanson

Figure 64 Cylinder pressure and AHRR of 3.5% DTBP addition to direct injected gasoline with 43% EGR (solid colored lines) and 50% EGR (dotted colored lines), and 9 (bar) IMEP compared to the gasoline/diesel operation of Hanson [47, 48] 43% EGR (black).

143 gasoline/3.5% DTBP gasoline (50% EGR) gasoline/3.5% DTBP gasoline (43% EGR) gasoline/diesel (R.M. Hanson)

Combustion duration 10-90 (deg)

12 11 10 9 8 6

8 6 4 2 0 -2

170 165

ETA thermal indicated

160 155 150 0.54

ISFC net (g/kW-hr)

CA50 (deg ATDC)

7 10

0.52 0.50 0.48 0.75 0.77 0.79 0.81 0.83 0.85 0.87 0.89 0.91 0.93 0.95

percent port fuel (%)

Figure 65 Combustion trends of 3.5% DTBP with 43 and 50% EGR and gasoline/diesel operation of Hanson [47, 48] with 43% EGR.

0.010

5

4.0

0.96

4 3 2 1 0

7

3.5 2.5 2.0

6

COV (%)

3.0

5 4 3

NOx(g/kW-hr)

1.0

0.2

2010 limit 0.1 0.0 0.76 0.78 0.80 0.82 0.84 0.86 0.88 0.90 0.92 0.94 0.96

percent port fuel (%)

Ex. CO2 (g/kW-hr)

1.5 0.3

0.99

0.97

0.000

26 22 18 14 10 6 2

1.00

0.98

HC (g/kW-hr)

PRR (bar/deg)

0.005

gasoline/3.5% DTBP gasoline (50% EGR) gasoline/3.5% DTBP gasoline (43% EGR) gasoline/diesel (R.M. Hanson)

η comb. (-)

2010 limit

0.015

550

2

540

1

CO (g/kW-hr)

0.020

PM (g/kW-hr)

gasoline/3.5% DTBP gasoline (50% EGR) gasoline/3.5% DTBP gasoline (43% EGR) gasoline/diesel (R.M. Hanson)

530 520 510 500 0.76 0.78 0.80 0.82 0.84 0.86 0.88 0.90 0.92 0.94 0.96

percent port fuel (%)

Figure 66 Emissions and combustion efficiency rends of 3.5% DTBP with 43 and 50% EGR and gasoline/diesel operation of Hanson [47, 48] with 43% EGR.

144 The combustion trends in Figure 65 are very interesting between the 43% and 50% EGR cases. As hypothesized, the additional EGR level resulted in decreasing the PRR, presented in Figure 66. PRR rate was decreased through later combustion phasing and increased combustion duration as noted by the delayed CA 50 timing and 10-90% burn duration trends. Even with additional intake temperature and reduced port fueling percentage, the combustion behavior is slower with the higher EGR level. Of particular interest the ISFC and thermal efficiency trends between the two EGR levels appear very similar in overall shape, but are offset from one another. The addition of EGR aided in lowering PRR, thus enabling a broader port fuel percentage sweep while maintaining low PRR. While at 43% EGR the PRR was above desirable levels at the 89% port fuel condition, preventing the investigation of reduced port fuel percentages. However, had lower port fuel percentages been investigated at the 43% EGR level, similar bowed ISFV and thermal efficiency trends would be expected as those demonstrated with 50% EGR, but shifted to higher port fuel percentages. Based on these trends, the 43% EGR minimum values of ISFC and maximum thermal efficiency would be expected to reside near or at the 89% port fuel condition. As seen from Figures 64-66, the only significant difference between operation at 43% and 50% EGR is that the pressure rise rate is decreased through changing the combustion duration and timing. Supporting this observation are the near identical emissions trends between the two EGR levels seen in Figure 66. Thus, the only significant trend difference seen between the

145 two EGR levels is the reduced PRR with the additional EGR. The ability of DTBP to successfully increase reactivity to the direct injected portion of a single low reactivity fuel stock shows application promise. The combustion and emissions trends of DTBP additized gasoline differ only slightly from those of diesel, and also show slight efficiency gains through decreased pumping work. Further optimization with EGR and DTBP percentage could lead to increased thermal efficiencies while maintaining pressure rise rates below 10 bar/deg or lead to low PRR operation at higher engine loads.

146

Chapter 6: Reactivity Reduced Fuel Reactivity Controlled Combustion 6.1

Reactivity reduction

The previous chapter discussed the combustion characteristics of fuel reactivity controlled combustion using port injected gasoline and direct injected gasoline with reactivity enhancement from the addition of a small percentage of DTBP. The results demonstrated that a percentage of DTBP addition above approximately 2% could effectively alter gasoline autoignition characteristics to behave similar to diesel fuel. To explore lower fuel reactivity effects, decreasing the port fuel’s reactivity was also investigated. Although lower reactivity (higher octane) non-oxygenated fuels exist – such as compressed natural gas or toluene- they are currently uncommon in their pure form at the end user level. An attractive and widely practiced method for increasing octane number is through ethanol addition. Primarily used with spark ignition engines, ethanol provides a high octane number (107 research octane number (RON)) [3], making it a lower reactivity fuel than most pump gasoline. For instance, E-85 (15% gasoline with 85% ethanol) has emerged within the marketplace as a high-octane fuel. Furthermore, 10% ethanol addition to gasoline has become common practice. 6.2

Ethanol Direct Injection Ignition Delay

Prior to experiments of fuel reactivity controlled combustion with port injection of ethanol, direct injection of ethanol was tested to examine its ignition sensitivity. Using identical operating conditions and the same single injection strategy as those of the direct-injection study in Table 13 with gasoline and DTBP in Chapter

147 5, ethanol was added to gasoline in various amounts. The ethanol used was E98 directly from the distiller (98 percent ethanol with 2% natural gasoline by volume). The specific fuel properties can be seen in Appendix B. Ethanol blends of the base ethanol fuel (E-98), 50 percent ethanol with gasoline (E-50), 25 percent ethanol with gasoline (E-25), 10 percent ethanol with gasoline (E-10), and 0 percent ethanol with gasoline (E-0) were tested. Ethanol direct-injection, compression-ignition combustion presents unique attributes. For example, although ethanol has a high RON, it also exhibits nontraditional properties compared to “conventional” fuels such as #2 ULSD and pump gasoline. Specifically, ethanol contains less energy with a lower heating value (LHV) of 26.9 MJ/kg (vs. ~43.3 MJ/kg for gasoline), has a richer air stoichiometric air fuel ratio of 9:1 (vs. ~14.6:1 for gasoline), and has a higher enthalpy of vaporization of 840 kJ/kg (vs. ~350 kJ/kg for gasoline) [3]. These three effects are of primary interest for ethanol use in piston cylinder engines. The decreased volume based fuel energy and richer fuel stoichiometry requires more fuel volume for a given engine load, decreasing volume-based fuel mileage. Beyond the benefits of higher octane number, the higher enthalpy of vaporization (charge cooling) of ethanol can be best utilized for reducing spark ignition engines propensity to knock, and to increase volumetric efficiency. However, the combination of these three effects creates a unique situation for direct-injected ethanol. Because of the reduced lower heating value of ethanol, for a given load the amount of fuel injected needs to be increased, but the fuel is less prone to auto

148 ignite and self cools (high octane number and enthalpy of vaporization). To compensate for the low fuel reactivity and high charge cooling characteristics of ethanol, engine intake air temperatures were elevated to phase combustion appropriately. Unfortunately, operation with E-98 was unrealizable at the desired load point of 3 (bar) IMEP. As injection duration was increased, so did charge cooling from the enthalpy of vaporization effect, which resulted in decreased fuel plume temperatures and increased ignition delay. As fueling increased the delay increased and eventually was severe enough to result in no combustion event. Intake temperature was increased to help to reduce the extended ignition delay, but the mechanical limits of the laboratory intake heaters were reached in doing so. Thus, E-98 operation at this condition was not possible; however, operation with the other ethanol blends was successful. The ignition delay trends of successful ethanol gasoline blends and gasoline both with and without DTBP are displayed in Figure 67, with the combustion properties displayed in Figure 68.

149 22

E-50 E-25 E-10 E-0 E-0 + 2% DTBP

number denotes intake temperature 100

EID [SOIC -CA 50] (deg)

20

85 70

18

45 55 30 35 40 55 60 45 30 70 35 40 50

16

14 55

12

40

50

40 45

55 45 50

10 -14

-12

-10

-8

-6

-4

-2

0

2

4

6

Injection Timing (deg ATDC)

Figure 67 Ignition delay trends of all gasoline-based fuels tested with the operating conditions shown Table 13, number indicates intake temperature. -16

-14

-12

-10

-8

-6

-4

-2

0

2

4

6 10 8

40 45 50 55

6

40 45 50 55

4

CA50 (deg ATDC)

2 0 50 55

50 45 55 40 30

30

35

40

60 70

35 45 55

45 55

40 45

50

40

220 45 55

210

30 35 40 50 60

ETA thermal indicated

30 35 40

200 70 40 45

45 55

40 45 50

50 55

190

55

180 170

0.46 0.44

100 C Intake

85 C 70 C 35 40 Intake Intake

40 45 45 50

50 55

50 55 30 35

70

55

40

0.40 0.38 -16

45 45

30

0.42

40

ISFC [G. eq.] net (g/kW-hr)

60 70 30 35 40 45 55

17 15 13 11 9 7 5

Combustion duration 10-90 (deg)

12 E-50 Gasoline E-25 Gasoline E-10 Gasoline E-0 Gasoline 50 E-0 2% DTBP Gasoline 45 55 30 35 40

55 60

-14

-12

-10

-8

-6

-4

-2

0

2

4

6

Injection Timing (deg ATDC)

Figure 68 Combustion trends of all gasoline-based fuels tested with the operating conditions shown in Table 13, number indicates intake temperature.

150 From Figures 67 and 68 it is seen that ethanol not only requires a significant intake temperature increase, but also simultaneously has increased ignition delay (EID). The longer ignition delay appears to have a minimal effect on the combustion duration and thermal efficiency, but operation below the user defined limit of 10 (bar/deg) does effect CA 50, exhibiting more advanced timings with higher ethanol percentages. The corresponding emissions of ethanol from Figure 69 shows that for all the gasoline-based fuels tested, all emissions and operation trends are similar except for CO, HC, and NOx, which are functions of temperature. 45

-6

-4

-2

0

2

4

-16 52 50

40

40 40

46

55 55

44

45 50

45 55

70 55 60

50 30 35 35 40 45 55 30 40 45 55 60 70

42

0.000

40 45 50 55

40 45 50 55

45 55 35 40 30

60 70 45

30 35 30 40

NOx (g/kW-hr)

10 8

100 C Intake

85 C Intake

30 35 30

40 45 50 55

40

45 50 55 70

45

6 35 40

4

55

60

40 45 50 55 40 45 50 55

70 C Intake

2 0 -16

40 45 50

50

55

4.0 3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0

55

-14

-12

-10

-8

-6

-4

-2

0

Injection Timing (deg ATDC)

2

4

6

-6

-4

-2

0

2

4

6

20.0 17.5 15.0 12.5 10.0 7.5 5.0 2.5 0.0

1.00 40 45 50 55 40 45 50 55

50 30 35 40 45 55 30

0.006

0.002

-8

1.05

0.008

0.004

-10

0.95

70

60 55

35 40 45

0.90

45 55 35

30

50

40

30

45 55 60 35 40 70

40 45 50 55 40 45 50 55

40 30

COV (%)

PM (g/kW-hr)

0.010

50

-12

E-50 Gasoline E-25 Gasoline E-10 Gasoline E-0 Gasoline E-0 Gasoline + 2% DTBP

HC (g/kW-hr)

35

48

40 45

45 50

30 30 35

-14

η comb. (-)

-8

45 50 55

30 35

20

40 30 35

700 650

35 30

70

45 40

40

50 55 45 50

10 55

50

40

30

550

60

0

85 C 70 C Intake Intake

100 C Intake

600

500 -16

40 45 55

CO (g/kW-hr)

-10

Ex. CO2 (g/kW-hr)

-12

E-50 Gasoline E-25 Gasoline E-10 Gasoline E-0 Gasoline E-10 Gasoline + 2% DTBP

IMEP (psi)

-14

45

40 45 55

50 40 45

55 60

70

40 45 50 55

55

35

-14

-12

-10

-8

-6

-4

-2

0

2

4

6

Injection Timing (deg ATDC)

Figure 69 Emissions and ignition delay for various gasoline based fuels. (Note that FID based HC values are uncorrected for fuel borne oxygen effects). Test conditions are presented in Table 13.

151 Ethanol displays the trend that the higher the ethanol percentage the higher the CO emissions. Although the HC emissions trends exhibit the opposite behavior, the HC emissions were acquired with a heated flame ionization detector (HFID) based measurement. This measurement is calibrated using the assumption that the incoming air has 21% oxygen. Although a valid assumption for a nonoxygenated fuel, oxygenated fuels invalidate this assumption, and the resulting measurements are thus confounded. Although FID analyzer-based pure ethanol HCCI combustion HC emissions by Sjöberg and Dec [68] were shown to be higher than those observed with PRF blends, a correction for oxygen percentage was required. This correction was not applied to the present direct-injected ethanol gasoline experiments because it assumes that all HC emissions result from unburned ethanol, which is not the case with ethanol blended fuels. Furthermore, empirical correlation studies such as those of Kar and Cheng [69] have been conducted using exhaust gas chromography-based measurements and correlated to HFIDbased response with oxygenated fuels. Such experiments are conducted to develop correction correlations for HFID-based measurements with fuels containing ethanol. Although such studies have demonstrated that a significant exhaust species fraction is actually pure ethanol with ethanol-gasoline blends as low as E-10, even pure E-100 operation results in roughly 50% oxygenated species in the exhaust HC emissions, with the remainder being similar to those of non-oxygenated fuels. The correlation developed by Kar and Cheng [69] is highly non-linear and results in an approximate 22% increase in HFID-based HC

152 emissions with E-100, a 15% difference for E-85, and only a 5% difference with E-50. Because of the small difference in HFID response with ethanol blends lower than E-50 or less, the correction was not applied in the present study. Regardless of the measured HC emissions, the increased CO emissions indicate that combustion does not reach completion or quenches, suggesting that HC emissions should correspond with this trend or remain unchanged. Additional supporting evidence for an increase in HC emissions with ethanol arises from the decreased NOx emissions. As discussed in Chapter 2, NOx is a strong function of temperature and forms from areas of locally high temperature (~1900 K or above). The NOx trends of Figure 69 indicate that as the percentage of ethanol is increased the charge cooling effects of the fuel quenches the local temperature and lowers NOx formation. This cooler local temperature would also be responsible for the increased CO and HC emissions from incomplete combustion within the fuel plume. The above results of direct-injection of gasoline with ethanol blends demonstrate that ethanol is clearly a reactivity-reducing species. Because compression ignition processes rely on early combustion stage chemistry, fuel effects become particularly important. Due to the wide availability of ethanol in the marketplace, the influence of port-injected ethanol on fuel reactivity controlled combustion is of much practical interest.

153 6.3

Fuel Reactivity Controlled Combustion with E-98 Ethanol and Diesel

The ethanol fuel used with fuel reactivity-controlled-combustion operation was identical to the prescribed ethanol above, and the properties are listed in Appendix B. As mentioned, port injected ethanol in place of gasoline has many inherent differences. Primarily, ethanol has a significantly lower heating value of 29.6 (kJ/kg) as compared to 43.22 (kJ/kg) for the tested gasoline. To account for the difference in heating value, fueling with E-98 port fueling was determined by matching fuel energy and not fuel mass. The difference in fuel heating value is a molecular effect that occurs from the alcohol group bond present in the molecule. Although the presence of oxygen also changes the fuel stoichiometry, for a given equivalence ratio, stoichiometric effects are negligible if the fuel energy is matched. However, a significant difference between ethanol and gasoline is that the enthalpy of vaporization of ethanol is much higher than that of gasoline. Therefore, for a given intake temperature the compression temperature history of ethanol will be cooler, and the engine volumetric efficiency can be higher due to the additional charge cooling provided by ethanol. These fuel effects were considered when the port injection ethanol experiments were conducted. 6.3.1

9 (bar) operation with E-98 port fuel

To explore the fundamental fuel and combustion behavior differences of ethanol, an intake temperature sweep was tested. With matched fuel energy to that of Hanson [47,48] and that of the DTBP gasoline tests of Table 16 in Chapter 5, the intake temperature was swept from 32°C to 44°C. To

examine fuel intake

154 sensitivity, for each intake temperature the fuel reactivity (port injected fuel amount) was altered to obtain constant peak main heat release timing. As seen in Figure 70, at the operating condition for a fixed peak heat release timing a 4°C increase in intake temperature is equivalent to a 1% decrease in port fuel percentage. The cylinder and AHRR trends of the sweep are presented in Figure 70, where the 89% port fueling case of Hanson [47, 48] is also plotted. Interestingly, unlike gasoline port injection, 9 (bar) IMEP operation with E-98 did not require EGR to maintain low PRR with load increases. Further insight and discussion for this requirement will be discussed in the sections to follow. 0% EGR (32°C) 0% EGR (36°C) 0% EGR (40°C) 0% EGR (44°C) 43% EGR (32°C) diesel/gasoline (R.M.Hanson)

100

0.035 0.030 0.025 0.020 0.015 0.010 0.005 0.000 -15

1.2 1.0 -13

-11

-9

-7

-5

Crank angle (deg)

80

1.4

NTC Behavior

0.8 77 76 75

60 89 baseline*

40

0.6 0.4

AHRR (kJ/deg)

Cylinder Pressure (bar)

120

AHRR (kJ/deg)

140

74 20 0 -25

0.2

% Port Fuel

-20

-15

-10

-5

0

5

10

Crank angle (deg)

15

20

0.0 25

*data from R.M.Hanson

Figure 70 Indicated cylinder pressure and AHRR for E-98 port fueling, with constant main heat release peak timing and 9 (bar) IMEP operation.

In post-processing the temperature sweep, it was found that the 44°C intake temperature condition resulted in nearly matched bulk gas temperature (within ±3

155 K) at -30° ATDC when compared to the 32°C intake te mperature of operation with port injection of gasoline. The nearly matched bulk gas temperatures just prior to combustion suggest that the increase in intake temperature from 32°C to 44°C nearly accounts for the engine system response to the higher enthalpy of vaporization effects with ethanol as compared to gasoline. The system response also includes the differences in gamma (ratio of specific heats) of the bulk gas through the presence of EGR with port injection of gasoline, and the lack of EGR with port injection of E-98 for operation. As noted in the legend of Figure 70, 9 (bar) operation with ethanol did not require the use of EGR to obtain low PRR combustion. The more triangular shape of the heat release of ethanol provides a steady transition into and through main heat release. As intake temperature is increased the triangular shape transitions closer to the more bell-shaped main heat release of gasoline/diesel operation. However, even with the nearly matched bulk gas temperature history with 44°C inta ke temperature, the E-98 fueled heat release event remains more triangular-shaped than the bell-shaped behavior of gasoline. To discount any combustion timing effects in the PRR and EGR requirements, port fueling operation with gasoline and both diesel and 1.75% DTBP were compared to operation with 44°C intake te mperature port fueling of E-98. Although the highest port fuel percentage with gasoline/diesel operation was 89% operation, with DTBP port fuel percentage operation up to 94% was tested. It was found that operation with 92% port gasoline and 1.75% DTBP resulted in similar peak heat release location to that of the 44°C intake

156 temperature E-98 condition. DTBP operation was compared because, direct comparison of E-98 to gasoline/diesel operation was not possible due to the combustion advance observed with the latest combustion timing with 89% port fuel gasoline/diesel. The pressure and AHRR traces of the comparison are shown in Figure 71.

E-98/diesel 44°C intake 0% EGR gasoline/diesel 32°C intake 43% EGR

100 80

AHRR (kJ/deg)

Pressure (bar)

120

gasoline/gasoline 1.75% DTBP 32°C intake 43% EGR NTC behavior

0.05 0.04 0.03 0.02 0.01 0.00 -25 -20 -15 -10 -5 Crank Angle (deg)

1.4 1.2 1.0

\p

0

89 baseline *

0.8

92

60

0.6

77

40

0.4

AHRR (kJ/deg)

140

% Port Fuel 20

0.2

0

0.0

-25 -20 -15 -10

-5

0

5

10

Crank Angle (deg)

15

20

25

* data from R.M.Hanson

Figure 71 Indicated cylinder pressure and AHRR of DTBP/gasoline, diesel/gasoline, and diesel/E-98 fuel reactivity-controlled combustion. Note the near matched expansion pressure confirms that fuel energy is matched for all cases.

As seen from the figure, operation with E-98 displays a smooth and earlier transition into the main heat release phase. These effects spread the heat

157 release event over more crank angles, decreasing the peak heat release rate magnitude, and enabling operation without the requirement of EGR to reduce PRR. This longer combustion duration behavior is different that that demonstrated by Sjöberg and Dec [68] where ethanol was found exhibit a more rapid heat release rate than PRF blends in traditional HCCI operation. Because both the results of Sjöberg and the present fuel reactivity combustion rely on chemical kinetic processes and not on flame speed affects, the increased flame speed of ethanol Egolfopoulus et al. [70] is not germane for analysis of the combustion process. As demonstrated in Chapter 3 and [66], reactivity-controlled-combustion initiates in areas of higher fuel reactivity and progresses to areas of lower reactivity. Thus, early/pre-combustion reactions are critical to the main combustion event. Based on this chemistry dependence, and the results shown in Figure 71, when compared to those of Sölberg and Dec [68]; the combustion process of port injected E-98 compared to that of gasoline in fuel reactivity controlled combustion is different than that of traditional HCCI. Although both combustion regimes are dependent on early chemistry kinetics for transition into the main heat release phase, the path to reach main heat release is different. With fuel reactivity controlled combustion the chemical kinetics are accelerated by reactive species produced during breakdown of the reactive fuel components, while traditional HCCI kinetics are purely dependent on the compressiontemperature-history-induced breakdown of the single fuel. This inherent

158 difference results in differences in the overall combustion chemistry, and in the nature of the combustion process. As noted, operation with E-98 port fuel required an increase in the reactive fuel percentage. Computations by Shi [71], suggest that ethanol attacks important free radicals, like the hydroperoxide radical, during the early stages of ethanol combustion. His predictions are supported by rapid compression machine (RCM) optical experiments by Hashimoto [72], which demonstrated that n-heptane NTC behavior was delayed with diminished luminosity when mixed with 30% ethanol by mass. Hashimoto states that these observations are consistent with predictions using chemical mechanisms developed by Curan et al. [73] with the ethanol mechanism of Marinov et al. [74] which suggest that ethanol attacks the hydroperoxide radical (OH) to form ethoxy (CH3CH2O·) and 1-hydroxy-1-ethyl (CH3CH·OH) molecules, instead of attacking n-heptane to produce heptyl radicals (CH3CH·C5H11). This is also supported by Shi [71], where HO2 chemistry is suggested to be very important in laminar flame combustion. Thus, the early cycle ethanol attack on radicals delays/alters the transition into the critical H2O2 loop chemistry, from the reactive fuel component Ando et al. [29]. The attack by ethanol on radicals results in slowed overall fuel decomposition through alteration of the fuel chemistry kinetic path as compared to gasoline. Similar results with non-oxygenated traditional fuels have been demonstrated by Shibata [31, 32] where inhibitor effects to the early fuel breakdown chemistry can dramatically alter combustion behavior. The early chemistry inhibitor effect of ethanol requires an increase in the reactive fuel

159 (diesel) quantity. This is needed to over-supply the radical pool to overcome the loss of reactive species from ethanol attack. Although these ethanol chemistry effects required the use of more diesel fuel, Figure 70 demonstrates that intake temperature increase can be substituted for increased reactive fuel quantity. This relation shows that there are competing kinetic effects. The fuel chemistry kinetic differences account for the dramatically different reactive component fueling requirements between port fueling of gasoline and E-98 in reactivity controlled combustion. Ethanol’s attack on important combustion intermediates slows the combustion chemistry, and enables engine operation without EGR to obtain low PRR. If EGR were required, additional main heat release combustion delay, would result, and additional operating parameters like fuel reactivity or intake temperature would need to be adjusted to account for the later combustion phasing. Although testing with EGR and ethanol was conducted, at the time of testing the intake temperature effects were not properly recognized, and the engine was operated with 32°C intake temperature (r esulting in bulk gas temperature of 32°C with only 59% port fuel). This testing also consisted of operation without EGR while sweeping port fueling percentage. Figure 72 depicts the indicated pressure and AHRR behavior of E-98 with EGR addition and 59% port fuel.

160 E-98/diesel 32°C intake (59% port fuel energy) gasoline/diesel 32°C intake (89% port fuel energy)

140

80

0.05 0.04 0.03 0.02 0.01 0.00 -25

1.2 1.0

-20 -15 -10 -5 Crank Angle (deg)

\p

0

43 baseline *

0.8

0 6 15

60

0.6 25

28

40 20

0.4 % EGR

0

AHRR (kJ/deg)

100

NTC behavior AHRR (kJ/deg)

Pressure (bar)

120

1.4

0.2 0.0

-25 -20 -15 -10

-5

0

5

10

Crank Angle (deg)

15

20

25

* data from R.M.Hanson

Figure 72 Indicated cylinder pressure and AHRR of E-98 EGR sweep at 9 (bar) IMEP, for the tested conditions shown in Table 16.

If the combustion timing was desired to be held constant throughout the EGR sweep, the fuel reactivity gradient would need to change to accelerate the fuel kinetics. Furthermore, accelerated kinetics can also make the heat release event more rapid, as demonstrated with gasoline compression ignition in Hanson et al. [9]. A similar combustion 10-90% duration trend with fuel reactivity-controlled combustion can be seen in Figure 73 where constant main heat release timing with E-98 operation without EGR exhibits longer burn durations than with EGR. From the results, EGR clearly delays the main heat release, and operation above 28% EGR with 32°C intake temperature was not possib le.

161 In addition to sweeping EGR with the 32°C intake p ort temperature, the port fuel percentage was also swept. The combustion phasing and performance trends of all ethanol sweeps (port fuel and EGR percentages at 32°C, and intake temperature without EGR) are overlaid with results of Hanson [47, 48]

and

operation with 1.75% DTBP addition to direct injected gasoline presented in the

E-98/ diesel 0% EGR, intake temp sweep E-98/ diesel EGR 32 [C] intake temp gasoline/gasoline 1.75 % DTBP, 43% EGR gasoline/diesel, 43% EGR

Combustion duration 10-90 (deg)

previous section, in Figure 73.

16

Reactive Fuel +

14 12

8

CA50 (deg ATDC)

EGR + 9 8 7 6 5 4 3 2 1 0 -1

6 Reactive Fuel +

EGR +

170 165 Reactive Fuel +

160

ETA thermal indicated

EGR + 155 150

ISFC net (g/kW-hr)

10

0.54 Reactive Fuel +

0.52 0.50 EGR + 0.56

0.60

0.64

0.68 0.80

0.84

0.88

0.92

0.96

percent port fuel energy (%)

Figure 73 Combustion performance and timing with E-98/diesel, gasoline/diesel and gasoline/gasoline + 1.75% DTBP. Note that ISFC is presented on a gasoline equivalent basis

With similar CA 50 timings, ethanol clearly extends the combustion duration. Regardless of the tested conditions, E-98 consistently exhibited lower thermal efficiencies and higher gasoline equivalent specific fuel consumption. The

162 difference indicates that the fuel is not utilized as well. For example, Table 18 depicts the relative cycle work and efficiencies of the cases varies and fuels presented in Figure 73. Table 18 Mechanical work breakdown for the fuels shown in Figure 73. Direct Injection Reactivity Fuel Diesel 1.75% DTBP Diesel Port Fuel Gasoline Gasoline E-98 EGR (%) 43 43 0 Intake Temperature (°C) 32 32 36 Port Fuel energy Percentage (%) 89 89 75 Indicated Compression pressure (bar) -8.83 -8.65 -8.69 Indicated Expansion pressure (bar)

18.02

18.06

17.85

Indicated Gross Power (kW)

24.28

24.87

22.88

Indicated Gross Thermal Efficiency (%)

56.2

57.3

54.8

Note that the compression work of E-98 is only slightly above that of gasoline DTBP, but the expansion work is lower. Lower expansion work results from either a difference in fueling, or combustion efficiency. Since the fuel energy was matched for these three cases, the decreased expansion work only can arise from lower combustion efficiency. The combustion losses are apparent in the deceased indicated thermal efficiency. To measure combustion losses of with ethanol, one must correct for fuel borne oxygen interaction with HFID-based HC measurements. As discussed above, this is not a linear trend and different correction techniques have been researched [68, 69]. To calculate combustion efficiency, both CO and HC emissions must be known. The trend in CO was obtained by the exhaust gas analyzers and is seen in Figure 74. However, HC emissions were not analyzed.

163 With premixed combustion HC and CO emissions have been shown to occur primarily from crevice flows [3]. Because a majority of the premixed fuel consists of shorter (light) hydrocarbons, the premixed HC emission trends are assumed to follow a similar distribution to that observed by Koci et al. [75], and were thus assumed to track with the CO emissions. At the operating conditions plotted in Figure 74 the exhaust CO emissions are similar, particularly at the lower port fueling percentages of E-98. Thus, little difference is assumed to occur in the combustion losses. However, when the port fueling percentage is increased there are additional CO emissions. This could occur from more ethanol becoming trapped in the crevice flows. The added ethanol also will provide additional charge cooling, possibly decreasing cylinder wall temperatures. Both of these effects would result in increased CO and HC emissions through incomplete combustion near cylinder boundary surfaces. As the port fuel energy percent is increase (green line), the CO emissions increase as compared to operation with ethanol with decreased port fuel percentage (purple line). This logic is thought to be universal for the tested conditions with operation of port injected ethanol. Also note: that little difference in CO emissions are observed between 32°C intake temperature and 44°C intake temperature with E-98, suggesting that intake temperature has a minimal effect on combustion efficiency.

164

0.01

5

HC (g/kW-hr)

0.00

Reactive Fuel +

EGR + 6.0 5.5 5.0 4.5 4.0 3.5 3.0 2.5 2.0 1.5 1.0

Reactive Fuel +

EGR + 0.8

NOx (g/kW-hr)

0.94

0.6 Reactive Fuel +

2010 limit

0.4 0.2 EGR + 0.0 0.56 0.60

0.64

0.68 0.80

0.84

percent port fuel (%)

0.88

0.92

0.96

0.92

4 3 2 1

13 11 9 7 Reactive Fuel +

5

EGR +

Ex. CO2 (g/kW-hr)

14 12 10 8 6 4 2

0.96

COV (%)

PRR (bar/deg)

EGR +

0.98

η comb. (-)

0.02

1.00

PM (g/kW-hr)

0.03

2010 limit

Reactive Fuel +

E-98/ diesel 0% EGR, intake temp sweep E-98/ diesel EGR 32 [C] intake temp gasoline/gasoline 1.75 % DTBP, 43% EGR gasoline/diesel, 43% EGR

3

550 530

CO (g/kW-hr)

E-98/ diesel 0% EGR, intake temp sweep E-98/ diesel EGR 32 [C] intake temp gasoline/gasoline 1.75 % DTBP, 43% EGR gasoline/diesel, 43% EGR

Reactive Fuel +

510 490 470

EGR + 450 0.56 0.60

0.64

0.68 0.80

0.84

0.88

0.92

0.96

percent port fuel energy (%)

Figure 74 Emissions trends of E-98/diesel, gasoline/gasoline + 1.75% DTBP, and gasoline/diesel operation [47, 48] HC emissions data with E98 not shown. Besides the CO trends with E-98, another interesting emissions trend seen in Figure 74 is the turn-up in NOx and PM observed with ethanol operation. In analyzing the higher NOx and PM emissions with ethanol, the reactive fuel percentage must be considered. As previously mentioned, operation with ethanol requires additional reactive fuel percentage due to ethanol attack of the early stage combustion radicals. This, in turn, results in areas where the reactive fuel resides becoming more fuel rich. As seen in the Phi-T plot of Sun [13], Figure 2, the additional reactive fuel percentage changes the local reactivity and can result in increased NOx and PM emissions. Figure 74 clearly displays this trend. Only when either the reactive fuel percent is decreased or mixing time increases from EGR addition, do NOx and PM emissions meet 2010 EPA mandate levels. This

165 trend demonstrates that the port fuel percent should be as high as possible to avoid rich hot areas that promote PM and NOx production. Conversely, higher port fueling percentage results in higher CO and assumed HC emissions. Of course different direct-injection fuel strategies could be explored to reduce this effect. The present research shows that ethanol is very beneficial in reducing the EGR and air-handling requirements of fuel-reactivity-controlled combustion. Nine (bar) IMEP operation was realized without EGR with low PRR. EGR tolerance was tested, and ethanol was found to have similar EGR tolerance as gasoline. However, ethanol was found to have slightly higher NOx and PM emissions. These were shown to be a function of diesel fuel fraction, and not due to the ethanol. Further exploration with different intake temperatures, injection timings, and perhaps EGR levels could be investigated to determine where ethanol operation can be further optimized for high efficiency and below 2010 EPA emissions.

166

Chapter 7: Conclusions 7.1

Summary of Investigations Engine experiments to further understand fuel reactivity controlled

combustion were conducted. Initially optical diagnostics were used to understand the staged ignition process of the combustion regime. Then alternative fuels with both increased and decreased reactivity were examined. Conclusions from each of the three experimental campaigns were made, as follows. 7.1.1

Experimental Conclusions The present crank angle resolved optical diagnostics of natural thermal

emission from FTIR-based broadband spectra in the mid IR range demonstrated that combustion initiates globally from regions containing the reactive fuel component of dual fuel operation. Once initiated, combustion continues in areas of high local fuel reactivity. However, areas of lower fuel reactivity demonstrate a pause between the initial NTC reactions and their high temperature heat release phase. These combustion conclusions were drawn from analysis of aldehyde and fuel reaction extents inferred from the measurements. The experimental measurements were also compared to KIVA-CFD predictions, and demonstrated good overall agreement in both timing and trends of the reaction extents. The measurements demonstrated that regardless of the local reactivity, gasoline and diesel fuels are steadily consumed during the high temperature heat release phase, while the computations made with PRF blends to represent the dual fuel blends predict that the blended individual fuels are consumed at different rates.

167 After the optical diagnostic testing, the effects of fuel reactivity were further examined, through the use of a peroxide-based cetane improver, DTBP. By this means the reactivity of gasoline could be enhanced. The results demonstrated that DTBP percentages above approximately 2% by volume are sufficient to provide near identical ignition behavior to that of the tested #2 ULSD. However, due to distillation differences, the direct injected fuel pressure with DTBP gasoline could be reduced to 400 (bar) from 800 bar. It was observed that DTBP plus gasoline exhibited significantly smaller and earlier NTC behavior than diesel fuel. The decreased early cycle heat release decreased compression work, and enabled an approximately 1% gain in net thermal efficiency. For constant CA 50 timing approximately 1% more port fuel was required for DTBP percentages above 1.75%. Also DTBP addition to gasoline

exhibited

nearly

identical

emissions

to

those

obtained

with

gasoline/diesel dual fueling. DTBP was found to have similar EGR tolerance to that of diesel. The findings demonstrate that relatively small amounts of cetane improver can successfully be used to increase the fuel reactivity of the directinjected fuel fraction, thus enabling use of a single low reactivity fuel stock. Lastly, experiments with decreased port injected fuel reactivity were conducted. The experiments used diesel for the direct-injected fuel, and E-98 ethanol port injected fuel. Operation with port injection of ethanol demonstrated that there were significant differences due to chemical phenomena. Primarily, ethanol blended with diesel fuel virtually eliminates NTC behavior, requiring additional diesel fuel percentage to maintain combustion phasing. Once high

168 temperature heat release was initiated, a slow progression through the heat release was observed, with a pronounced triangular shape, compared to the more bell shaped heat release obtained with gasoline/diesel fueling. The extended and slowly transitioning heat release event allowed low PRR operation at 9 (bar) IMEP without EGR. For a given combustion timing, ethanol also required approximately 12°C higher intake temperatu res to compensate for the enthalpy of vaporization, EGR, and fuel heating value differences. Also a 4°C difference in intake temperature was found to be equivalent to a 1% difference in port fuel percentage. Operation with ethanol demonstrated that increased NOx and particulate emissions resulted if the direct injection events were not appropriately

advanced.

Thermal

efficiency

was

slightly

reduced

by

approximately 1% with direct-injected diesel and port-injected ethanol, as compared to port injected gasoline. 7.1.2

Suggestions for Future Work

The present optical experiments used natural thermal emissions measurements. More advanced absorption or laser-based optics could be used to further quantify the combustion process. The measurement range could be further extended into the IR for observation of CO emissions. Optical measurements with additional reference fuels, cetane improvers, and oxygenated fuels could also be explored. This data would be particularly useful for model validation studies. Also, optical measurements at higher loads and equivalence ratios could be examined. Fuel reactivity was examined, but operation at peak engine loads was not demonstrated. Operation at higher loads and lower and higher engine speeds

169 could be examined to investigate the operational bounds of the fuel-reactivity controlled combustion. For the experiments conducted thus far a relatively high compression ratio has been employed. Lower compression ratios could be investigated to allow more operational flexibility in intake and EGR temperatures, and also in EGR rate requirements. Lastly, multi-cylinder and transient experiments could be explored to investigate cylinder wall temperature effects, and to demonstrate transient control of the combustion regime. The present work has helped to further understand the combustion process and fuel effects of fuel reactivity-controlled combustion. This combustion regime is very promising as a means to achieve very high thermal efficiency with very low NOx and particulate emissions. Further research to better understand and develop this combustion regime, whish shows promise for reducing fuel consumption, engine emissions, and engine cost is suggested.

170

References [1]

Stone, Richard, “Introduction to Internal Combustion Engines” Third Edition, SAE, 1999

[2]

De Nevers, Noel, “Air Pollution Control Engineering”, McGraw Hill, 2000

[3]

Heywood, J. B., “Internal Combustion Engine Fundamentals,” McGraw-Hill Inc., 1988

[4]

Eng J.A., "Characterization of pressure waves in HCCI combustion", SAE Technical Paper 2002-01-2859, 2002

[5]

Dec, J.E., “A Conceptual Model of DI Diesel Combustion Based on Laser Sheet Imaging”, SAE Paper 970873, 1997.

[6]

Kamimoto T., Yokota H., Kobayashi H., “Effect of high pressure injection on soot formation processes in a rapid compression machine to simulate diesel flames”, SAE Paper 871610, 1987.

[7]

Nabier J.D., Sibers D.L.,“Effects of gas density and vaporization on penetration and dispersion of diesel sprays”, SAE paper 960034, 1996.

[8]

Kalghatgi, G.T., Risberg, P. and Ångström, H.-E., “Partially Pre-Mixed Auto Ignition of Gasoline to Attain Low Smoke and Low NOx at High Load in a Compression Ignition Engine and Comparison with a Diesel Fuel”, SAE paper 2007-01-0006, 2007.

[9]

Hanson, R.M., Splitter, D.A., Reitz, R.D., “Operating a Heavy-Duty DirectInjection Compression-Ignition Engine with Gasoline for Low Emissions”, SAE paper 2009-01-1442, 2009.

[10]

Puntambekar A.K., “Combustion and Spray Visualization in a High-Speed Direct Injected Engine”, M.S. Thesis, University of Wisconsin – Madison, 2009.

[11]

Turns, S.R., “An introduction to combustion: Concepts and Applications”, New York, McGraw-Hill inc., 1995.

[12]

T. Kamimoto and M. Bae, "High combustion temperature for the reduction of particulate in diesel engines", SAE Technical Paper 880423, 1988.

[13]

Y. Sun, "Diesel combustion optimization and emissions reduction using adaptive injection strategies (AIS) with improved numerical models", PhD Thesis, University of Wisconsin-Madison, 2007.

171

[14]

Hardy, W., “An Experimental Investigation of Advanced Diesel Combustion Strategies for Emissions Reductions in a Heavy-Duty Diesel Engine at High Speed and Medium Load”, M.S. Thesis, University of Wisconsin – Madison, 2005.

[15]

K. Akihama, Y. Takatori, K. Inagaki, S. Sasaki and A. M. Dean, "Mechanism of the smokeless rich diesel combustion by reducing temperature", SAE Technical Paper 2001-01-0655, 2001.

[16]

Kimura, S., Aoki, O., Ogawa, H., Muranaka, S. and Enomoto, Y., “A New Combustion Concept for Ultra-Clean and High-Efficiency Small DI Diesel Engines,” SAE Paper 1999-01-3681,1999

[17]

Hasegawa R., Yanagihara H., “HCCI combustion in a DI Diesel Engine”, SAE 2003-01-0745, 2003.

[18]

M. Christensen, A. Hultqvist and B. Johansson, "Demonstrating the multifuel capability of a homogeneous charge compression ignition engine with variable compression ratio", SAE Technical Paper 1999-01-3679, 1999.

[19]

Koci, C.P., “Multiple-Event Fuel Injection and Emission Investigations in a Low Temperature Combustion Regime Using a Small Bore Diesel Engine”, M.S. Thesis, University of Wisconsin – Madison, 2008.

[20]

Onishi S, Jo S.H., Shoda K., Jo P.D., Kato S.: ”Active ThermoAtmosphere Combustion (ATAC) – A New Combustion Process for Internal Combustion Engines”, SAE Technical Paper 790501, 1979

[21]

Najt P.M., Foster D.E., “Compression-ignited homogeneous charge combustion” SAE Technical Paper 830264, 1983

[22]

Thring, R.H., “Homogeneous Charge Compression Ignition (Hcci) Engines” (SAE Technical Paper 892068, 1989.

[23]

P. W. Bessonette, C. H. Schleyer, K. P. Duffy, W. L. Hardy and M. P. Liechty, "Effects of fuel property changes on heavy-duty HCCI combustion", SAE Technical Paper 2007-01-0191, 2007.

[24]

Dec J.E., Hwang W., Sjöberg M., “Potential of thermal stratification and combustion retard for reducing pressure-rise rates in HCCI engines, based on multi-zone modeling and experiments” SAE Technical paper, 2006-01-1518, 2006

[25]

Sjöberg M., Dec J.E., Cernansky N.P., “Potential of Thermal Stratification and Combustion Retard for Reducing Pressure-Rise Rates in HCCI

172 Engines, Based on Multi-Zone Modeling and Experiments”, SAE 2005-010113, 2005. [26]

Iverson R.J., “The Effects of Intake Charge Stratification on HCCI Combustion”, M.S. Thesis, University of Wisconsin – Madison, 2004.

[27]

Sjöberg M.,Dec J.E., Hwang W., “Thermodynamic and Chemical Effects of EGR and Its Constituents on HCCI Autoignition”, SAE 2007-01-0207, 2007.

[28]

Herold R.E., Krasselt J.M., Foster D.E., Ghandhi J.B., Reuss D.L., Najt P.M., “Investigations into the Effects of Thermal and Compositional Stratification on HCCI Combustion - Part II: Optical Engine Results”, SAE Technical Paper 2009-01-1106, 2009.

[29]

Ando H., Sakai Y., and Kuwahara K., "Universal rule of hydrocarbon oxidation", SAE Technical Paper 2009-01-0948, 2009.

[30]

Kalghatgi G.T., “Auto-Ignition Quality of Practical Fuels and Implications for Fuel Requirements of Future SI and HCCI Engines”, SAE 2005-010239, 2005.

[31]

Shibata G., Urushihara T., “Auto-Ignition Characteristics of Hydrocarbons and Development of HCCI Fuel Index”, SAE 2007-01-0220, 2007.

[32]

Shibata G., Oyama K., Urushihara T., and Nakano T., “Correlation of Low Temperature Heat Release With Fuel Composition and HCCI Engine Combustion”, SAE 2005-01-0138, 2005.

[33]

Bunting B.G., Wildman C.B., Szybist J.P., Lewis S., and Storey J., “Fuel chemistry and cetane effects on diesel homogeneous charge compression ignition performance, combustion, and emissions”, Journal of Engine Research vol 8, no 1, pp 15-27, 2007.

[34]

Kalghatgi G.T., Risberg P., and Ångström H., “Advantages of Fuels with High Resistance to Auto-ignition in Late-injection, Low-Temperature, Compression Ignition Combustion” SAE 2006-01-3385, 2006.

[35]

Westbrook, C. K., “Chemical Kinetics of Hydrocarbon Ignition in Practical Combustion Systems", Proceedings of the Combustion Institute, Volume 28, 2000/pp. 1563–1577.

[36]

Thompson, A.A., Lambert, S.W., Mulqueen, S.C., “Prediction and precision of cetane number improver response equations”, SAE 972901, 1997.

173 [37]

“Guidance on Quantifying NOx Benefits for Cetane Improvement Programs for Use in SIPs and Transportation Conformity” EPA420-B-04005, June 2004.

[38]

Tanaka S., Ayala F., Keck J.C., Heywood J.B., “Two-stage ignition in HCCI combustion and HCCI control by fuels and additives”, Combustion and Flame 132 (2003) 219–239.

[39]

Eng J.A., Leppard, W.R., and Sloane, T.M., “The Effect of Di-Tertiary Butyl Peroxide (DTBP) Addition to Gasoline on HCCI Combustion” SAE 200301-3170, 2003.

[40]

Ickes, A.M., “Fuel Property Impact on a Premixed Diesel Combustion Mode”, PhD. Thesis, University of Michigan –Ann Arbor, 2009.

[41]

Kumar S., Stanton D.W., Fang H., Gustafson R.J., Frazier T.R., Bunting B.G., Xu Y., and Wolf L.R., “The Effect of Diesel Fuel Properties on Engine-out Emissions and Fuel Efficiency at Mid-Load Conditions” SAE 2009-01-2697, 2009.

[42]

Oxley J.C., Smith J.L., Rogers E., Ye W., Aradi A.A., and Henly T.J., “Fuel Combustion Additives: A Study of Their Thermal Stabilities and Decomposition Pathways”, Energy & Fuels 2000, 14, 1252-1264.

[43]

Higgens B, Siebers D., Mueller C, and AradiI A., “Effects of an IgnitionEnhancing, Diesel-Fuel Additive on Diesel-Spray Evaporation, Mixing, Ignition, and Combustion”, Twenty-Seventh Symposium (International) on Combustion/The Combustion Institute, 1998/pp. 1873–1880.

[44]

Alam M., Goto S., Sugiyama K., Kajiwara M., and Mori M., Konno M., Motohashi M., and Oyama K., “Performance and Emissions of a DI Diesel Engine Operated with LPG and Ignition Improving Additives”, SAE 200101-3680, 2001.

[45]

Inagaki K., Fuyuto T., Nishikawa K., Nakakita K. and Sakata I., "Dual-fuel PCI combustion controlled by in-cylinder stratification of ignitability", SAE Technical Paper 2006-01-0028, 2006.

[46]

Kokjohn S.L., Hanson R.M., Splitter D.A., and Reitz R.D., "Experiments and Modeling of Dual-Fuel HCCI and PCCI Combustion Using In-Cylinder Fuel Blending", SAE Technical Paper 2009-01-2647", 2009.

[47]

Hanson R.M., Splitter D.A., Kokjohn S.L., and Reitz R.D., "An Experimental Investigation of Fuel Reactivity Controlled PCCI Combustion in a Heavy-Duty Engine", SAE Technical Paper 2010-01-0864, 2010.

174 [48]

Hanson, R.M., “Experimental Investigation of Fuel Effects on Low Temperature Combustion in a Heavy Duty Compression-Ignition Engine”, M.S. Thesis, University of Wisconsin – Madison, 2009.

[49]

Reitz R.D, Kokjohn S.L., Hanson R.M. Splitter D.A., “Engine Combustion Control via Fuel Reactivity Stratification”, United States Patent Application P100054, 2010

[50]

Staples L.R., “Experimental Investigation of Diesel Engine Size Scaling Parameters”, , M. S. Thesis, University of Wisconsin – Madison, 2008.

[51]

Hruby E., “An Experimental Investigation of a Homogeneous Charge Compression Ignition Engine using Low Pressure Injection and Diesel Fuel” M.S. Thesis, University of Wisconsin-Madison 2003.

[52]

Weninger E.D., Private communication, University of Wisconsin, Madison 2007.

[53]

Liechty, M. P., “Optimization of Heavy-Duty Diesel Engine Operating Parameters at High Speed and Medium Load Using µ-Genetic Algorithms”, M. S. Thesis, University of Wisconsin – Madison, 2004.

[54]

Tess M., Private Communication, University of Wisconsin, Madison 2009

[55]

Nevin R.M., “PCCI Investigation Using Variable Intake Valve Closing in a Heavy Duty Diesel Engine”, M. S. Thesis, University of Wisconsin – Madison, 2006.

[56]

Lee, S., “Investigation of Two Low Emissions Strategies for Diesel Engines; Premixed Charge Compression Ignition (PCCI) and Stoichiometric Combustion” ”, Ph.D. dissertation, University of WisconsinMadison, 2006.

[57]

Thiel M. “Application of Automated Experiments to the Optimization of a Direct-Injected Diesel Engine for the Simultaneous Reduction of NOx and Particulate Emissions” M.S. Thesis, University of Wisconsin-Madison, 2001.

[58]

Rein K.D., Sanders S.T., S. L. Lowery S.L., Jiang E.Y., Workman Jr. J.J., "In-Cylinder Fourier-transform infrared spectroscopy", Measurement Science and Technology, 19 (2008) 043001 (5pp).

[59]

Rein K.D., Bartula R.J., Sanders S.T, "Interferometric Techniques for Crank-Angle Resolved Measurements of Gas Spectra in Engines", SAE Technical Paper 2009-01-0863, 2009.

175 [60]

Summers D.P., Abrams M.C., Brault J.W., "Fourier Transform Spectrometry", Academic Press, 2001.

[61]

Williams R.D., Liechty M.P., Private Communication, Caterpillar Inc., Peoria, Illinois, 2009.

[62]

Klingbiel, A.E., “Mid-IR laser absorption diagnostics for hydrocarbon vapor diagnostics in harsh environments", PhD. dissertation, Stanford University, 2007

[63]

Rothman L.S., Gordon I.E., Barbe A., Benner D.Chris, Bernath P.F., Birk M., Boudon V., Brown L.R., Campargue A., Champion J.-P., Chance K., Coudert L.H., Dana V., Devi V.M., Fally S., Flaud J.-M., Gamache R.R., Goldman A., Jacquemart D., Kleiner I., Lacome N., Lafferty W.J., Mandin J.-Y., Massie S.T., Mikhailenko S.N., Miller C.E., Moazzen-Ahmadi N., Naumenko O.V., Nikitin A.V., Orphal J., Perevalov V.I., Perrin A., PredoiCross A., Rinsland C.P., Rotger M., Šimečková M., Smith M.A.H., Sung K., Tashkun S.A., Tennyson J., Toth R.A., Vandaele A.C., and Vander Auwera J. "The HITRAN 2008 molecular spectroscopic database", Journal of Quantitative Spectroscopy & Radiative Transfer 110 (2009) 533–572.

[64]

NIST Website http://webbook.nist.gov/chemistry/form-ser.html

[65]

Meyers A. M., "High Spectral Resolution Emission Thermometry for Combustion Applications", M.S. Thesis University of Wisconsin Madison, 2005

[66]

Splitter D.A, Kokjohn S.L., Rein K.D., Hanson R.D., Sanders S.T., Reitz R.D., “An Optical Investigation of Ignition Process in Fuel Reactivity Controlled PCCI Combustion”, SAE Technical Paper 2010-01-0345, 2010.

[67]

Kranendonk, L. A., Caswell, A. W., Meyers, A. M., and Sanders, S. " Wavelength-Agile Laser Sensors for Measuring Gas Properties in Engines", SAE Technical Paper 2003-01-1116, 2003

[68]

Sjöberg M, Dec J.E., “Ethanol Autoignition Characteristics and HCCI Performance over Wide Ranges of Engine Speed, Load and Boost “, SAE Technical Paper 2010-01-0338, 2010

[69]

Kar, K. Cheng, W.K., “Speciated Engine-Out Organic Gas Emissions from a PFI-SI Engine Operating on Ethanol/Gasoline Mixtures”, SAE Technical Paper 2009-01-2673, 2009.

[70]

Egolfopoulus F.E., DU D.X., Law C.K., “A study on Ethanol Oxidations Kinetics in Laminar Premixed Flames, Flow Reactors, and Shock Tubes”

176 24th symposium (international) on Combustion/ the Combustion Institute,1992, pp 833-841. [71]

Shi Y, “Optimization of a Compression Ignition Engine Fueled With Diesel and Gasoline-Like Fuels”, PhD dissertation, University of WisconsinMadison,2009

[72]

Hashimoto K., “Inhibition Effect of Ethanol on Homogeneous Charge Compression Ignition of Heptane”, SAE Technical Paper 2008-01-2504, 2008.

[73]

Curran, H.J., Gaffuri, P., Pitz, W.J., Westbrook, C.K., “A Comprehensive Study of N-heptane Oxidation”, Combustion and Flame 144, 149-177, (1998).

[74]

Marinov, N.M., “A Detailed Chemical Kinetic Model for High Temperature Ethanol Oxidation”, International Journal of Chemical Kinetics 31, 183-220 (1999).

[75]

Koci C.P., Ra Y, Krieger R., Andrie M., Foster D.E., Siewert R.M., Durrett R.P., Ekoto I, Miles P.C., “Detailed Unburned Hydrocarbon Investigations in a Highly-Dilute Diesel Low Temperature Combustion Regime”, SAE Technical Paper 2009-01-0928, 2009

177

Appendix A:

Part Drawings and Prints

The following prints are for The Optical Head as designed by Caterpillar. Note that this cylinder head has additional firedeck material for threading of combustion visualization adaptors. Also prints for the common rail adaptor are included with the supporting clamps. The common rail adaptor will allow Bosch GDI and common rail injectors, as well as Denso style common rail injectors to be used in the SCOTE. The head must be fitted with HEUI 315 adaptors (part numbers ERP 7871 and ERP 7872). When using the adaptor a crush seal must be inserted at the injector adaptor interface. Crush seal thickness can be altered to adjust the injector spray targeting if desired.

178

179

180

181

182

183

184

185

186

187

188

189

190

191

192

193

194

195

Appendix B:

Fuel Properties

196

197

198

199

200

Appendix C:

Engine Data

Optical Experiments Data Indicated Cylinder Pressure and Apparent Heat Release Rate Optical Tests Measured Cylinder Pressure and AHRR

60

NTC behavior 0.05 0.04 0.03 0.02 0.01 0.00 -25 -20 -15 -10

0.4 -5

0

Crank Angle (deg)

0.3

40

0.2

20

0.1

0

AHRR (kJ/deg)

Pressure (bar)

80

0.5 AHRR (kJ/deg)

100

0.0

-25 -20 -15 -10

-5

0

5

10

15

20

25

Crank Angle (deg)

Engine Performance and Exhaust Emissions Data Run # Fuel/Fuel % Main Fuel E Blend % Pilot Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC] Dp/Dtheta Fuel Rail Pressure [bar]

g1_c3 3.00 G/D 0.32 0.00 0.68 650.00 300.00 5000.00 -35.00 -37.00 -58.00 -320.00 4.13 800.00

201

brake emissions Exhaust_Temp_(C) NOx_(g/kW-hr) HC_(g/kW-hr) NOx_+_HC_(g/kW-hr) PM_(g/kW-hr) CO_(g/kw-hr) Intake_CO2_(g/kW-hr) Exhaust_CO2_(g/kW-hr) SOF_(%) AVL_Equivalence_Ratio AVL_C_(mg/m^3)

195.00 0.02 7.40 7.42 0.01 22.83 12.15 885.00 0.00 0.00 0.74

emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_(g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f]

0.09 26.74 26.83 0.03 82.54 3199.45 43.91

net indicated emissions NOx_(g/ikW-hr) HC_(g/ikW-hr) NOx_+_HC_(g/ikW-hr) PM_(g/ikW-hr) CO_(g/ikw-hr) Intake_CO2_(g/ikW-hr) Exhaust_CO2_(g/ikW-hr) Engine_Speed_(RPM) Intake_Pressure_(psi) Exhaust_Pressure_(psi) Intake_Temp_(C) Intake_Flowrate_(kg/min) Upstream_Orifice_Press_(psi) EGR_(%) EGR_Temp_(C) HP_Fuel_Flowrate_(lb/hr) BSFC_(g/kW-hr) Brake_Torque_(ft-lbs) Brake_Power_(kW) LP_Fuel_Flowrate_(lb/hr) VFD_Setpoint(Hz) BSFC_(g/kW-hr)

0.02 4.97 4.98 0.01 15.33 8.16 594.36 1300.00 20.00 21.00 32.00 2.38 64.54 0.00 25.00 1.65 90.98 46.00 8.49 3.54 0.00 276.61

202 BMEP calc [Psi] BMEP calc [Bar]

46.57 3.21

Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr]

2.35 0.04 5.19

ETA thermal indicated ETA Gross Indicated

0.46 0.50

Carbon mass out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_Deviation_of_IMEP(psi) COV_of_IMEP(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur) ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC gross [g/kw-hr] Indicated Power [kw] Gross power [kw] eta-comb (heywood 4.96) combustion efficiency Carbon Based Equivalence Ratio phi A/F (A/F)s

2166.16 2056.63 1.05 70.92 1.68 2.37 5.16 4.78 -5.49 -4.25 0.50 3.00 5.27 7.01 9.52 10.76 185.77 185.77 172.32 12.64 13.63 0.05 0.95

0.25 57.14 14.48

203

Optical In-Cylinder Data

Location B (Squish) Raw Spectra -19 deg ATDC

-17 deg ATDC

Location B

0.04

0.05

Measured Spectra

Measured Spectra

0.05

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

0.04

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm) -15 deg ATDC

Wavelength (nm) -13 deg ATDC

Location B

0.04

0.05

Measured Spectra

Measured Spectra

0.05

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

Location B

Location B

0.04

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

204 -11 deg ATDC

-9 deg ATDC

Location B

0.04

0.05

Measured Spectra

Measured Spectra

0.05

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

0.04

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

-7 deg ATDC

Wavelength (nm)

-5 deg ATDC

Location B

0.04

0.05

Measured Spectra

Measured Spectra

0.05

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

-3 deg ATDC

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

-1 deg ATDC

Location B

0.05

Measured Spectra

Measured Spectra

Wavelength (nm)

0.04

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

Location B

0.04

Wavelength (nm)

0.05

Location B

Location B

0.04

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

205 1 deg ATDC

3 deg ATDC

Location B

0.04

0.05

Measured Spectra

Measured Spectra

0.05

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

0.04

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm) 5 deg ATDC

Wavelength (nm) 7 deg ATDC

Location B

0.04

0.05

Measured Spectra

Measured Spectra

0.05

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

9 deg ATDC

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

16 deg ATDC

Location B

0.05

Measured Spectra

Measured Spectra

Wavelength (nm)

0.04

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

Location B

0.04

Wavelength (nm)

0.05

Location B

Location B

0.04

0.03

0.02

0.01

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

206

Location A (Bowl) Raw Spectra -19 deg ATDC

-17 deg ATDC

0.20

0.20 Location A

Measured Spectra

Measured Spectra

Location A

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

Wavelength (nm)

-15 deg ATDC

-13 deg ATDC

0.20

0.20 Location A

Measured Spectra

Measured Spectra

Location A

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

207 -11 deg ATDC

-9 deg ATDC

0.20

0.20 Location A

Measured Spectra

Measured Spectra

Location A

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

Wavelength (nm)

-7 deg ATDC

-5 deg ATDC

0.20

0.20 Location A

Measured Spectra

Measured Spectra

Location A

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

Wavelength (nm) -1 deg ATDC

-3 deg ATDC 0.20

0.20

Location A

Measured Spectra

Measured Spectra

Location A

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

208 1 deg ATDC

3 deg ATDC

0.20

0.20 Location A

Measured Spectra

Measured Spectra

Location A

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

Wavelength (nm)

5 deg ATDC

7 deg ATDC

0.20

0.20 Location A

Measured Spectra

Measured Spectra

Location A

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

Wavelength (nm)

9 deg ATDC

18 deg ATDC

0.20

0.20 Location A

Measured Spectra

Measured Spectra

Location A

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

209 5 deg ATDC

7 deg ATDC

0.20

0.20 Location A

Measured Spectra

Measured Spectra

Location A

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

0.15

0.10

0.05

0.00 2400 2600 2800 3000 3200 3400 3600 3800

Wavelength (nm)

210

3 (bar) IMEPn DI tests E-0 Indicated Cylinder Pressure and Apparent Heat Release Rate

0.8

80

0.8

60

0.6

60

0.6

Evap. Cooling

0.4

0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

Evap. Cooling 0.04 0.02 0.00

Crank Angle (deg)

0

0.0

-10

-5

0

5

10

15

20

0

25

0.0

-20

-15

-10

Crank Angle (deg)

10

15

20

25

3 (bar) IMEPn Testing, E-0, Run5

0.8

80

0.8

60

0.6

60

0.6

Evap. Cooling

0.06

0.4

0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

0.0

-10

-5

0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

0

5

10

15

20

0

25

0.0

-20

-15

-10

Crank Angle (deg)

-5

0

5

10

15

20

25

Crank Angle (deg)

3 (bar) IMEPn Testing, E-0, Run6

3 (bar) IMEPn Testing, E-0, Run7 100

80

0.8

80

0.8

60

0.6

60

0.6

0.06

Evap. Cooling

0.4

0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

Crank Angle (deg)

0.0

-15

-10

-5

0.06

NTC behavior

0.4

0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

Crank Angle (deg)

0

-20

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

20

AHRR (kJ/deg)

40

Pressure (bar)

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

100

0.4

0.04

Crank Angle (deg)

0

-15

Evap. Cooling

0.06

Crank Angle (deg)

-20

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

40

Pressure (bar)

80

AHRR (kJ/deg)

100

Cylinder Pressure AHRR

AHRR (kJ/deg)

Pressure (bar)

5

1.0

20

Pressure (bar)

0

Crank Angle (deg)

3 (bar) IMEPn Testing, E-0, Run4 100

-5

AHRR (kJ/deg)

-15

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

Crank Angle (deg)

-20

0.4

0.06

0

5

10

Crank Angle (deg)

15

20

25

0

-20

0.0

-15

-10

-5

0

5

10

Crank Angle (deg)

15

20

25

AHRR (kJ/deg)

20

0.06

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

40

Pressure (bar)

80

AHRR (kJ/deg)

100

AHRR (kJ/deg)

Pressure (bar)

3 (bar) IMEPn Testing, E-0, Run3 1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

3 (bar) IMEPn Testing, E-0, Run2 100

211 3 (bar) IMEPn Testing, E-0, Run8

3 (bar) IMEPn Testing, E-0, Run9

0.8

80

0.8

60

0.6

60

0.6

0.4

0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

Evap. Cooling 0.04 0.02 0.00

Crank Angle (deg)

0

0.0

-10

-5

0

5

10

15

20

0

25

0.0

-20

-15

-10

Crank Angle (deg)

10

15

20

25

3 (bar) IMEPn Testing, E-0, Run11

0.8

80

0.8

60

0.6

60

0.6

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

Evap. Cooling 0.04 0.02 0.00

0.0

-10

-5

0

5

10

15

20

0

25

0.0

-20

-15

-10

Crank Angle (deg)

-5

0

5

10

15

20

25

Crank Angle (deg)

3 (bar) IMEPn Testing, E-0, Run12

3 (bar) IMEPn Testing, E-0, Run13 100

80

0.8

80

0.8

60

0.6

60

0.6

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

Crank Angle (deg)

0.0

-15

-10

-5

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

Crank Angle (deg)

0

-20

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

20

AHRR (kJ/deg)

40

Pressure (bar)

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

100

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

Crank Angle (deg)

0

-15

0.4

0.06

Crank Angle (deg)

-20

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

40

Pressure (bar)

80

AHRR (kJ/deg)

100

Cylinder Pressure AHRR

AHRR (kJ/deg)

Pressure (bar)

5

1.0

20

Pressure (bar)

0

Crank Angle (deg)

3 (bar) IMEPn Testing, E-0, Run10 100

-5

AHRR (kJ/deg)

-15

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

Crank Angle (deg)

-20

0.4

0.06

0

5

10

Crank Angle (deg)

15

20

25

0

-20

0.0

-15

-10

-5

0

5

10

Crank Angle (deg)

15

20

25

AHRR (kJ/deg)

20

Evap. Cooling 0.06

AHRR (kJ/deg)

40

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

80

Pressure (bar)

100

AHRR (kJ/deg)

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

Pressure (bar)

100

212 3 (bar) IMEPn Testing, E-0, Run14

80

0.8

60

0.6

40

20

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

AHRR (kJ/deg)

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

Pressure (bar)

100

Crank Angle (deg)

0

-20

0.0

-15

-10

-5

0

5

10

15

20

25

Crank Angle (deg)

E-0 Engine Performance and Exhaust Emissions Data (Runs 2-7) g1_c2 2.00 E0 1.00 0.00 0.00

g1_c3 3.00 E0 1.00 0.00 0.00

g1_c4 4.00 E0 1.00 0.00 0.00

g1_c5 5.00 E0 1.00 0.00 0.00

g2_c2 6.00 E0 1.00 0.00 0.00

g2_c3 7.00 E0 1.00 0.00 0.00

860.00

860.00

860.00

860.00

860.00

860.00

0.00 0.00

0.00 0.00

0.00 0.00

0.00 0.00

0.00 0.00

0.00 0.00

-4.00

-5.00

-6.00

-5.00

-6.00

-7.00

-1.50

-2.50

-3.50

-2.50

-3.50

-4.50

Dp/Dtheta

3.74

3.76

5.70

3.64

3.65

5.12

Fuel Rail Pressure [bar]

1000

1000

1000

1000

1000

1000

Run # Fuel/Fuel % Main Fuel E Blend % Pilot Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

brake emissions Exhaust_Tem p_(C)

213 NOx_(g/kWhr) HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr) CO_(g/kw-hr) Intake_CO2_ (g/kW-hr) Exhaust_CO 2_(g/kW-hr) SOF_(%) AVL_Equival ence_Ratio AVL_C_(mg/ m^3)

emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_( g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f] net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr) Intake_CO2_ (g/ikW-hr) Exhaust_CO 2_(g/ikW-hr)

16.61 21.05

14.42 25.29

16.47 22.34

13.75 35.63

14.23 29.17

15.17 23.57

37.65

39.72

38.81

49.38

43.40

38.74

0.00 26.48

0.00 32.61

0.00 27.32

0.00 42.93

0.00 36.29

0.00 29.59

36.36

40.34

39.13

47.19

42.94

38.93

1182 0.00

1156 0.00

1142 0.00

1292 0.00

1191 0.00

1098 0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.04

0.07

0.05

0.05

0.08

0.09

42.58 53.95

36.98 64.84

44.20 59.94

30.46 78.97

34.85 71.46

40.71 63.26

96.53 0.00 67.88

101.82 0.01 83.61

104.15 0.00 73.32

109.43 0.00 95.14

106.32 0.01 88.90

103.97 0.01 79.42

3031

2965

3064

2865

2918

2948

93.22

103.43

105.01

104.58

105.21

104.48

8.43

7.40

8.45

6.41

6.98

7.84

10.68

12.98

11.46

16.60

14.32

12.18

19.10

20.38

19.90

23.01

21.30

20.01

0.00

0.00

0.00

0.00

0.00

0.00

13.43

16.73

14.01

20.00

17.81

15.29

18.44

20.70

20.07

21.99

21.08

20.11

599.78

593.43

585.76

602.50

584.79

567.52

214 Engine_Spee d_(RPM) Intake_Press ure_(psi) Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr] ETA thermal indicated ETA Gross Indicated Carbon mass out/hr carbon mass in/hr

1300

1300

1300

1300

1300

1300

30.00

30.00

30.00

30.00

30.00

30.00

31.30

31.20

31.40

31.30

31.30

31.30

70.00

60.00

55.00

55.00

50.00

45.00

3.17

3.29

3.31

3.31

3.33

3.33

0.00

0.76

0.76

0.80

0.79

0.76

3.50

3.50

3.50

3.50

3.50

3.50

343.17

335.61

346.08

383.02

377.55

346.65

22.00

22.00

23.00

19.00

21.00

23.00

4.06

4.06

4.25

3.51

3.88

4.25

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

390.08

390.08

372.64

451.20

408.17

372.64

22.27

22.27

23.28

19.23

21.26

23.28

1.54

1.54

1.61

1.33

1.47

1.61

1.58

1.58

1.58

1.58

1.58

1.58

0.03

0.03

0.03

0.03

0.03

0.03

3.50

3.50

3.50

3.50

3.50

3.50

0.42

0.42

0.44

0.40

0.42

0.43

0.49

0.49

0.51

0.47

0.49

0.51

1406

1440

1468

1409

1434

1427

1375

1376

1376

1378

1377

1376

215 Carbon balance IMEP_(psi) (endaq) Standard_De viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur) ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC Gross EID [CA, SOI-CA50] Indicated Power [kw] Indicated Power [kw] eta-comb (heywood 4.96)

Carbon Based Equivalence Ratio phi A/F (A/F)s

1.02

1.05

1.07

1.02

1.04

1.04

44.89

44.40

45.35

42.33

44.37

46.14

1.15

1.20

0.91

1.53

1.28

0.97

2.56

2.70

2.02

3.62

2.90

2.09

3.55

3.52

3.68

3.38

3.54

3.66

3.03 11.77 12.04 13.58 15.13 17.78 20.61

2.99 11.07 11.35 13.02 14.77 17.13 20.02

3.13 9.32 9.58 11.02 12.11 14.88 18.08

2.85 11.60 12.03 13.80 16.01 19.08 21.04

2.99 10.31 10.61 12.32 14.13 16.79 19.35

3.11 8.85 9.27 10.76 12.07 14.81 17.86

5.74

5.78

5.30

7.05

6.18

5.54

6.01

6.06

5.56

7.48

6.48

5.96

197.87

200.13

191.12

210.25

200.35

192.48

197.87 168.64

200.13 170.04

191.12 162.72

210.25 177.21

200.35 169.26

192.48 163.51

15.08

15.52

14.52

16.30

15.82

15.26

8.00

7.91

8.29

7.53

7.90

8.23

9.39

9.31

9.73

8.94

9.36

9.69

0.07 0.93

0.08 0.92

0.08 0.92

0.10 0.90

0.09 0.91

0.08 0.92

0.13 116.22 14.61

0.12 121.28 14.61

0.12 119.37 14.61

0.12 123.82 14.61

0.12 123.60 14.61

0.12 123.77 14.61

216 g2_c5 8.00 E0 1.00 0.00 0.00

g3_c2 9.00 E0 1.00 0.00 0.00

g3_c3 10.00 E0 1.00 0.00 0.00

g3_c4 11.00 E0 1.00 0.00 0.00

g3_c5 12.00 E0 1.00 0.00 0.00

g1_c4 13.00 E0 1.00 0.00 0.00

g1_c4 14.00 E0 1.00 0.00 0.00

860.00

860.00

860.00

860.00

860.00

860.00

860.00

0.00 0.00

0.00 0.00

0.00 0.00

0.00 0.00

0.00 0.00

0.00 0.00

0.00 0.00

-6.00

-7.00

-8.00

-8.00

-9.00

-9.00

-10.00

-3.50

-4.50

-5.50

-5.50

-6.50

-6.50

-7.50

Dp/Dtheta

3.68

3.68

6.97

3.95

7.53

4.86

7.39

Fuel Rail Pressure [bar]

1000

1000

1000

1000

1000

1000

1000

12.08 36.26

16.23 20.23

16.23 20.23

11.40 27.22

11.60 25.20

16.24 20.70

16.24 20.70

48.34

36.46

36.46

38.62

36.79

36.94

36.94

0.00 46.31

0.00 24.30

0.00 24.30

0.00 41.96

0.00 44.75

0.00 31.19

0.00 31.19

47.86

36.95

36.95

42.43

44.07

41.01

41.01

1255 0.00

1049 0.00

1049 0.00

1123 0.00

1144 0.00

1124 0.00

1124 0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.07

0.06

0.06

0.08

0.10

0.11

0.11

Run # Fuel/Fuel % Main Fuel E Blend % Pilot Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

brake emissions Exhaust_Tem p_(C) NOx_(g/kWhr) HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr) CO_(g/kw-hr) Intake_CO2_ (g/kW-hr) Exhaust_CO 2_(g/kW-hr) SOF_(%) AVL_Equival ence_Ratio AVL_C_(mg/ m^3)

217

emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_( g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f] net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr) Intake_CO2_ (g/ikW-hr) Exhaust_CO 2_(g/ikW-hr) Engine_Spee d_(RPM) Intake_Press ure_(psi) Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kW-

26.78 80.36

47.34 59.02

47.34 59.02

29.22 69.79

28.41 61.73

43.59 55.54

43.59 55.54

107.14 0.01 102.63

106.37 0.01 70.89

106.37 0.01 70.89

99.02 0.01 107.56

90.14 0.01 109.63

99.13 0.01 83.70

99.13 0.01 83.70

2782

3062

3062

2879

2804

3016

3016

106.07

107.78

107.78

108.76

107.97

110.06

110.06

5.64

9.61

8.84

5.85

5.30

8.58

8.18

16.93

11.98

11.03

13.98

11.51

10.93

10.43

22.56

21.59

19.87

19.83

16.80

19.51

18.61

0.00

0.00

0.00

0.00

0.00

0.00

0.00

21.62

14.39

13.24

21.54

20.44

16.47

15.72

22.34

21.88

20.13

21.78

20.13

21.66

20.67

586.11

621.82

572.13

576.70

522.77

593.73

566.51

1300

1300

1300

1300

1300

1300

1300

30.00

30.00

30.00

30.00

30.00

30.00

30.00

31.50

31.40

31.40

31.30

31.30

31.40

31.40

45.00

40.00

40.00

35.00

35.00

30.00

30.00

3.37

3.45

3.45

3.53

3.49

3.55

3.55

0.83

0.75

0.75

0.75

0.78

0.75

0.75

3.50 425.39

3.50 170.49

3.50 170.49

3.50 366.10

3.50 374.56

3.50 343.24

3.50 343.24

218 hr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr] ETA thermal indicated ETA Gross Indicated Carbon mass out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_De viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC]

19.00

25.00

25.00

22.00

21.00

23.00

23.00

3.51

4.62

4.62

4.06

3.88

4.25

4.25

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

451.20

342.79

342.79

390.08

408.17

372.64

372.64

19.23

25.31

25.31

22.27

21.26

23.28

23.28

1.33

1.74

1.74

1.54

1.47

1.61

1.61

1.58

1.58

1.58

1.58

1.58

1.58

1.58

0.03

0.03

0.03

0.03

0.03

0.03

0.03

3.50

3.50

3.50

3.50

3.50

3.50

3.50

0.40

0.41

0.45

0.42

0.45

0.42

0.44

0.47

0.49

0.52

0.49

0.52

0.50

0.52

1398

1464

1464

1427

1383

1447

1447

1378

1375

1375

1377

1377

1376

1376

1.01

1.06

1.06

1.04

1.00

1.05

1.05

42.25

43.80

47.52

44.38

47.61

45.14

47.26

1.73

1.24

0.72

1.17

0.73

1.01

0.76

4.09

2.83

1.52

2.64

1.54

2.24

1.61

3.40

3.50

3.76

3.55

3.77

3.61

3.75

2.84 10.84 11.29 13.10 15.29 18.35 20.31

2.95 9.76 10.08 11.84 13.80 16.60 19.04

3.21 7.52 7.77 9.04 10.04 12.81 16.06

2.99 8.80 9.26 10.87 12.76 15.55 18.09

3.21 6.59 6.86 8.26 9.13 12.07 15.57

3.04 7.79 8.10 9.77 11.26 13.87 16.80

3.19 6.05 6.32 7.63 8.76 11.61 15.11

219 CA 10-90 (comb dur) CA 5-90 (comb dur) ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC Gross EID [CA, SOI-CA50] Indicated Power [kw] Indicated Power [kw] eta-comb (heywood 4.96)

Carbon Based Equivalence Ratio phi A/F (A/F)s

7.06

6.52

5.03

6.28

5.21

5.78

5.29

7.51

6.84

5.29

6.75

5.48

6.08

5.57

210.62

203.01

186.79

200.25

186.42

196.80

187.78

210.62 176.41

203.01 171.29

186.79 159.38

200.25 168.88

186.42 158.92

196.80 165.94

187.78 159.75

16.60

16.34

14.54

16.37

14.76

16.27

15.13

7.52

7.80

8.48

7.91

8.50

8.05

8.43

8.98

9.25

9.94

9.38

9.97

9.54

9.91

0.10 0.90

0.08 0.92

0.08 0.92

0.09 0.91

0.09 0.91

0.07 0.93

0.07 0.93

0.11 128.77 14.61

0.12 124.73 14.61

0.12 124.73 14.61

0.11 131.65 14.61

0.11 133.93 14.61

0.11 129.81 14.61

0.11 134.29 14.61

220

E-0 + 2% DTBP Indicated Cylinder Pressure and Apparent Heat Release Rate 3 (bar) IMEPn Testing, E-0 + 2% DTBP, Run3

0.8

80

0.8

60

0.6

60

0.6

0.4

0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

Evap. Cooling 0.04 0.02 0.00

Crank Angle (deg)

0

0.0

-10

-5

0

5

10

15

20

0

25

0.0

-20

-15

-10

Crank Angle (deg)

10

15

20

25

3 (bar) IMEPn Testing, E-0 + 2% DTBP, Run5

0.8

80

0.8

60

0.6

60

0.6

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

Evap. Cooling 0.04 0.02 0.00

0.0

-10

-5

0

5

10

15

20

0

25

0.0

-20

-15

-10

Crank Angle (deg)

-5

0

5

10

15

20

25

Crank Angle (deg)

3 (bar) IMEPn Testing, E-0 + 2% DTBP, Run6

3 (bar) IMEPn Testing, E-0 + 2% DTBP, Run7 100

80

0.8

80

0.8

60

0.6

60

0.6

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

Crank Angle (deg)

0.0

-15

-10

-5

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

Crank Angle (deg)

0

-20

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

20

AHRR (kJ/deg)

40

Pressure (bar)

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

100

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

Crank Angle (deg)

0

-15

0.4

0.06

Crank Angle (deg)

-20

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

40

Pressure (bar)

80

AHRR (kJ/deg)

100

Cylinder Pressure AHRR

AHRR (kJ/deg)

Pressure (bar)

5

1.0

20

Pressure (bar)

0

Crank Angle (deg)

3 (bar) IMEPn Testing, E-0 + 2% DTBP, Run4 100

-5

AHRR (kJ/deg)

-15

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

Crank Angle (deg)

-20

0.4

0.06

0

5

10

Crank Angle (deg)

15

20

25

0

-20

0.0

-15

-10

-5

0

5

10

Crank Angle (deg)

15

20

25

AHRR (kJ/deg)

20

Evap. Cooling 0.06

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

40

Pressure (bar)

80

AHRR (kJ/deg)

100

AHRR (kJ/deg)

Pressure (bar)

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

3 (bar) IMEPn Testing, E-0 + 2% DTBP, Run2 100

221

0.8

80

0.8

60

0.6

60

0.6

20

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

Crank Angle (deg)

0.0

-15

-10

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

Crank Angle (deg)

0

-20

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

40

Pressure (bar)

80

AHRR (kJ/deg)

100

AHRR (kJ/deg)

Pressure (bar)

3 (bar) IMEPn Testing, E-0 + 2% DTBP, Run9 1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

3 (bar) IMEPn Testing, E-0 + 2% DTBP, Run8 100

-5

0

5

10

15

20

0

25

0.0

-20

-15

-10

Crank Angle (deg)

-5

0

5

10

15

20

25

Crank Angle (deg)

E-0 Engine Performance and Exhaust Emissions Data (Runs 2-9) g1_c2 2.00 E0_2% DTBP

g1_c3 3.00 E0_2% DTBP

g1_c4 4.00 E0_2% DTBP

g1_c5 5.00 E0_2% DTBP

g2_c2 6.00 E0_2% DTBP

g2_c3 7.00 E0_2% DTBP

g2_c5 8.00 E0_2% DTBP

g3_c2 9.00 E0_2% DTBP

1.00 0.00 0.00

1.00 0.00 0.00

1.00 0.00 0.00

1.00 0.00 0.00

1.00 0.00 0.00

1.00 0.00 0.00

1.00 0.00 0.00

1.00 0.00 0.00

Main SOI1 Dur us

860.00

860.00

860.00

860.00

860.00

860.00

860.00

860.00

Main SOI2 Dur us

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

Pilot Dur us

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

Main SOI1 [ATDC]

-2.00

-1.00

-1.00

0.00

0.00

1.00

1.00

2.00

0.50

1.50

1.50

2.50

2.50

3.50

3.50

4.50

10.32

4.82

5.72

4.47

5.12

3.87

4.35

3.77

Run # Fuel/Fuel % Main Fuel E Blend % Pilot

Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC] Dp/Dtheta

222

Fuel Rail Pressure [bar]

1000

1000

1000

1000

1000

1000

1000

1000

brake emissions NOx_(g/kWhr)

15.21

14.11

14.10

13.76

14.30

12.90

13.85

13.03

9.72

10.47

9.40

10.64

9.89

11.38

10.23

10.80

24.92

24.58

23.49

24.40

24.19

24.28

24.08

28.44

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

CO_(g/kw-hr)

6.74

9.61

5.04

8.87

5.28

10.37

5.97

17.84

Intake_CO2_ (g/kW-hr)

36.14

36.43

35.47

37.98

38.08

39.12

38.97

45.45

Exhaust_CO 2_(g/kW-hr) SOF_(%)

1239 0.00

1220 0.00

1181 0.00

1216 0.00

1236 0.00

1206 0.00

1231 0.00

1364 0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.09

0.08

0.10

0.08

0.08

0.04

0.11

0.05

NOx_(g/kg-f)

40.73

37.79

39.53

36.85

38.30

34.55

37.09

30.18

HC_(g/kg-f) NOx_+_HC_( g/kg-f)

26.02

28.05

26.34

28.49

26.49

30.49

27.40

25.01

66.75

65.84

65.87

65.34

64.80

65.04

64.48

65.85

PM_(g/kg-f)

0.01

0.01

0.01

0.01

0.01

0.00

0.01

0.01

CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f]

18.06

25.74

14.14

23.75

14.13

27.76

15.99

41.31

3319

3267

3311

3256

3311

3232

3298

3159

96.80

97.56

99.45

101.73

102.00

104.77

104.38

105.23

HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr)

AVL_Equival ence_Ratio AVL_C_(mg/ m^3)

emissions index emissions

223

net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr)

7.39

7.00

7.20

6.82

7.02

6.52

6.90

5.95

4.72

5.20

4.80

5.27

4.86

5.75

5.10

4.94

12.12

12.20

11.99

12.09

11.88

12.27

12.00

12.99

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

3.28

4.77

2.57

4.39

2.59

5.24

2.98

8.15

Intake_CO2_ (g/ikW-hr)

17.58

18.08

18.10

18.82

18.70

19.76

19.43

20.76

Exhaust_CO 2_(g/ikW-hr)

602

605

602

602

606

609

613

623

Engine_Spee d_(RPM)

1300

1300

1300

1301

1300

1300

1300

1300

Intake_Press ure_(psi)

30.00

30.00

30.00

30.00

30.00

30.00

30.00

30.00

31.40

31.40

31.40

31.40

31.50

31.30

31.40

31.40

70.00

60.00

55.00

55.00

50.00

45.00

45.00

40.00

3.42

3.42

3.38

3.37

3.33

3.33

3.29

3.29

82.59 0.00

82.64 0.00

81.40 0.00

81.40 0.00

80.06 0.00

80.38 0.58

78.91 0.59

79.26 0.64

3.35

3.35

3.35

3.35

3.35

3.35

3.35

3.35

371.12

366.69

370.33

349.20

363.65

374.86

358.41

431.64

NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr)

Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr)

224

Brake_Torqu e_(ft-lbs) Brake_Power _(kW)

22.00

22.00

23.00

22.00

22.00

22.00

22.00

19.00

4.06

4.06

4.25

4.06

4.06

4.06

4.06

3.51

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

7.00

7.00

7.00

7.00

7.00

7.00

373.36

373.36

356.67

373.36

373.36

373.36

373.36

431.86

22.27

22.27

23.28

22.27

22.27

22.27

22.27

19.23

1.54

1.54

1.61

1.54

1.54

1.54

1.54

1.33

1.52

1.52

1.52

1.52

1.52

1.52

1.52

1.52

0.03

0.03

0.03

0.03

0.03

0.03

0.03

0.03

3.35

3.35

3.35

3.35

3.35

3.35

3.35

3.35

ETA thermal indicated

0.46

0.45

0.46

0.45

0.45

0.44

0.45

0.42

ETA Gross Indicated

0.54

0.53

0.54

0.53

0.53

0.52

0.53

0.50

LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr]

Carbon mass out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_De viation_of_IM EP(psi) COV_of_IME P(pct)

1407

1410

1418

1398

1418

1400

1415

1370

1316

1316

1316

1317

1317

1317

1317

1319

1.07

1.07

1.08

1.06

1.08

1.06

1.07

1.04

46.80

45.88

46.68

45.96

46.37

45.10

45.68

43.15

0.60

0.76

0.65

0.72

0.58

0.77

0.62

1.24

1.28

1.65

1.39

1.56

1.25

1.72

1.36

2.87

225

Gross IMEP [bar] IMEP NET [bar]

3.71

3.64

3.71

3.65

3.67

3.58

3.62

3.45

3.16

3.09

3.15

3.10

3.13

3.04

3.08

2.91

CA5 [ATDC]

10.08

11.56

10.53

12.29

11.54

13.59

12.79

15.27

CA10 [ATDC]

10.55

11.84

10.78

12.57

11.79

14.03

13.06

15.76

CA50 [ATDC]

11.62

13.32

12.07

14.07

13.11

15.58

14.55

17.56

CA75 [ATDC]

12.77

14.53

13.39

15.29

14.57

16.86

15.78

19.03

CA90 [ATDC]

16.33

17.79

17.29

18.56

18.34

20.12

19.13

22.10

CA95 [ATDC]

19.86

21.28

20.57

21.83

21.60

23.57

22.54

25.06

CA 10-90 (comb dur)

5.78

5.96

6.52

5.98

6.55

6.09

6.06

6.35

CA 5-90 (comb dur)

6.25

6.24

6.77

6.27

6.80

6.53

6.34

6.84

ISFC net [g/kw-hr]

181.57

185.29

182.04

184.98

183.31

188.62

186.11

197.29

ISFC gasoline eq. [g/kw-hr]

181.57

185.29

182.04

184.98

183.31

188.62

186.11

197.29

ISFC gross [g/kw-hr]

11.12

11.82

10.57

11.57

10.61

12.08

11.05

13.06

Indicated Power [kw]

8.35

8.18

8.33

8.19

8.27

8.04

8.14

7.68

Gross power [kw]

9.80

9.63

9.81

9.64

9.71

9.46

9.58

9.13

eta-comb (heywood 4.96)

0.03

0.03

0.03

0.03

0.03

0.04

0.03

0.03

combustion efficiency

0.97

0.97

0.97

0.97

0.97

0.96

0.97

0.97

226 Carbon Based Equivalence Ratio phi A/F (A/F)s

0.12 125.85 14.61

0.12 126.95 14.61

0.12 124.55 14.61

0.12 125.72 14.61

0.12 123.01 14.61

0.12 124.73 14.61

0.12 121.65 14.61

0.12 125.61 14.61

227

3 (bar) IMEPn DI tests E-10 Indicated Cylinder Pressure and Apparent Heat Release Rate

0.8

80

0.8

60

0.6

60

0.6

20

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

40

20

Crank Angle (deg)

0.0

-15

-10

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -10-8 -6 -4 -2 0 2 4 6 8 1012

Crank Angle (deg)

0

-20

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

40

Pressure (bar)

80

AHRR (kJ/deg)

100

AHRR (kJ/deg)

Pressure (bar)

3 (bar) IMEPn Testing, E-10 Run3 1.0

Cylinder Pressure AHRR

-5

0

5

10

15

20

Crank Angle (deg)

25

0

-20

0.0

-15

-10

-5

0

5

10

15

20

Crank Angle (deg)

E-10 Engine Performance and Exhaust Emissions Data Run # Fuel/Fuel % Main Fuel E Blend % Pilot Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

g1_c2

g1_c3

2 E10/G90

3 E10/G90

1.00 0.10 0.00

1.00 0.10 0.00

870

870

0

0

0

0

-10

-9

-7.5

-6.5

25

AHRR (kJ/deg)

3 (bar) IMEPn Testing, E-10 Run2 100

228

Dp/Dtheta

4.79

3.63

Fuel Rail Pressure [bar]

1000

1000

brake emissions NOx_(g/kWhr)

17.05

15.08

HC_(g/kW-hr)

19.26

30.35

36.31

45.43

0.00

0.00

CO_(g/kw-hr)

44.49

69.62

Intake_CO2_ (g/kW-hr)

39.96

50.08

Exhaust_CO 2_(g/kW-hr) SOF_(%)

1273 0.00

1498 0.00

0.00

0.00

0.08

0.06

NOx_(g/kg-f)

40.89

28.92

HC_(g/kg-f) NOx_+_HC_( g/kg-f)

46.20

58.20

87.09

87.12

PM_(g/kg-f)

0.01

0.01

CO_(g/kg-f) Exh CO2 [g/kg-f]

106.71

133.50

3054

2873

NOx_+_HC_( g/kW-hr) PM_(g/kWhr)

AVL_Equival ence_Ratio AVL_C_(mg/ m^3) emissions index emissions

229 INT CO2 [g/kg-f]

95.83

96.03

8.109

5.932

9.163

11.937

17.271

17.869

0.002

0.001

21.163

27.381

Intake_CO2_ (g/ikW-hr)

19.005

19.697

Exhaust_CO 2_(g/ikW-hr)

605

589

Engine_Spee d_(RPM)

1300

1300

Intake_Press ure_(psi)

30

30

31.3

31.3

70

70

3.165

3.165

76.53 0

76.58 0.54

3.4

3.4

363.13

460.89

net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr)

Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr)

230

Brake_Torqu e_(ft-lbs) Brake_Power _(kW)

20

16

3.69

2.95

0

0

0

0

416.93

521.51

20.25

16.20

1.40

1.12

1.54

1.54

0.03

0.03

1.48

1.48

3.40

3.40

ETA thermal indicated

0.436

0.421

ETA Gross Indicated

0.516

0.500

LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow gasoline eq. [kg/hr] Fuel Flow [lbm/hr]

Carbon mass out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_De viation_of_IM EP(psi) COV_of_IME P(pct)

1366

1313

1285

1287

1.06

1.02

44.48

42.13

0.73

1.4

1.65

3.32

231 Gross IMEP [bar] IMEP NET [bar]

3.48

3.37

2.93

2.84

CA5 [ATDC]

8.35

9.58

CA10 [ATDC]

8.60

9.85

CA50 [ATDC]

10.08

11.59

CA75 [ATDC]

11.34

13.35

CA90 [ATDC]

13.85

16.06

CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur)

16.87

18.52

5.25

6.21

-3.09

-3.37

198.32

205.11

190.83

197.36

161.06

166.35

17.58

18.09

7.76 9.19

7.50 8.90

eta-comb (heywood 4.96)

0.070

0.088

combustion efficiency

0.929

0.911

0.114 125.8435 14.04585

0.114 129.6145 14.04585

ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC Gas eq. gross EID [CA, SOI-CA50] Indicated Power [kw] Gross power

Carbon Based Equivalence Ratio phi A/F (A/F)s

232

3 (bar) IMEPn DI tests E-25 Indicated Cylinder Pressure and Apparent Heat Release Rate 3 (bar) IMEPn Testing, E-25 Run2

3 (bar) IMEPn Testing, E-25 Run3

0.8

80

0.8

60

0.6

60

0.6

20

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -16-14-12-10-8 -6 -4 -2 0 2 4 6 8

40

20

AHRR (kJ/deg)

40

Crank Angle (deg)

0.0

-15

-10

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -16-14-12-10-8 -6 -4 -2 0 2 4 6 8

Crank Angle (deg)

0

-20

1.0

Cylinder Pressure AHRR

-5

0

5

10

15

20

Crank Angle (deg)

25

0

-20

0.0

-15

-10

-5

0

5

10

15

20

Crank Angle (deg)

E-25 Engine Performance and Exhaust Emissions Data g1_c2

g1_c3 2

3

Run # E25/G50

E25/G50

Fuel/Fuel 1.00

1.00

0.25

0.25

0.00

0.00

885

885

0

0

0

0

-12

-13

% Main Fuel E Blend % Pilot Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us

Main SOI1 [ATDC] Main SOI2 [ATDC]

25

AHRR (kJ/deg)

80

Pressure (bar)

100

AHRR (kJ/deg)

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

Pressure (bar)

100

233 -9.5

-10.5

Dp/Dtheta

4.46

6.84

Fuel Rail Pressure [bar]

1000

1000

5.59

13.08

22.92

18.90

28.51

31.98

0.00

0.00

76.82

42.31

Intake_CO2_ (g/kW-hr)

38.29

35.26

Exhaust_CO 2_(g/kW-hr)

1272

1248

0.00

0.00

0.00

0.00

0.01

0.04

11.24

29.22

Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

brake emissions NOx_(g/kWhr) HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr) CO_(g/kw-hr)

SOF_(%) AVL_Equival ence_Ratio AVL_C_(mg/ m^3)

emissions index emissions NOx_(g/kg-f)

234 46.07

42.23

57.31

71.45

0.00

0.00

154.43

94.53

2557

2789

76.97

78.79

2.507

6.252

10.280

9.035

NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr)

12.787

15.287

0.000

0.001

34.45

20.22

Intake_CO2_ (g/ikW-hr)

17.1

16.8

Exhaust_CO 2_(g/ikW-hr)

570

596

Engine_Spee d_(RPM)

1300

1300

Intake_Press ure_(psi)

30

30

Exhaust_Pre ssure_(psi) Intake_Temp _(C)

31.3

31.2

100

100

Intake_Flowr ate_(kg/min)

3.086

3.085

HC_(g/kg-f) NOx_+_HC_( g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f]

net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr)

235 Upstream_Or ifice_Press_( psi)

73.81

73.49

0

0

3.65

3.65

423.34

361.74

18

20

3.32

3.69

0

0

0

0

497.46

447.58

18.22

20.25

1.26

1.40

Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow gasoline eq. [kg/hr] Fuel Flow [lbm/hr]

1.65

1.65

0.03

0.03

1.50

1.50

3.65

3.65

ETA thermal indicated

0.412

0.429

0.49

0.51

EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar]

ETA Gross Indicated

236

Carbon mass out/hr carbon mass in/hr Carbon balance

IMEP_(psi) (endaq) Standard_De viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar]

1269

1359

1293

1293

0.98

1.05

41.76

42.99

1.02

0.72

2.45

1.68

3.33

3.44

2.80

2.92

9.02

4.76

9.28

5.28

10.63

6.79

11.86

7.56

14.33

8.35

17.29

13.76

5.05

3.07

-3.97

-1.70

223.11

213.95

202.05

193.76

169.88

164.30

20.13

17.29

7.40

7.72

CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur)

ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC Gas eq. gross EID [CA, SOI-CA50] Indicated Power [kw]

237 8.80

9.10

0.046

0.042

0.953

0.957

0.117

0.117

112.1

112.0

13.19

13.19

Gross power

eta-comb (heywood 4.96) combustion efficiency

Carbon Based Equivalence Ratio phi A/F (A/F)s

238

3 (bar) IMEPn DI tests E-50 Indicated Cylinder Pressure and Apparent Heat Release Rate

80

0.8

80

0.8

60

0.6

60

0.6

0.4

0.04 0.02 0.00

0.2

-0.02 -16-14-12-10-8 -6 -4 -2 0 2 4 6 8

40

20

Crank Angle (deg)

0.0

-15

-10

Evap. Cooling

0.4

0.06 0.04 0.02 0.00

0.2

-0.02 -16-14-12-10-8 -6 -4 -2 0 2 4 6 8

Crank Angle (deg)

0

-20

1.0

Cylinder Pressure AHRR

AHRR (kJ/deg)

20

Evap. Cooling 0.06

AHRR (kJ/deg)

40

Pressure (bar)

100

AHRR (kJ/deg)

Pressure (bar)

3 (bar) IMEPn Testing, E-50 Run3 1.0

Cylinder Pressure AHRR

-5

0

5

10

15

20

0

25

-20

Crank Angle (deg)

0.0

-15

-10

-5

0

5

10

15

20

Crank Angle (deg)

3 (bar) IMEPn Testing, E-50 Run4

100

1.0

80

0.8

60 40

0.6

20

Evap. Cooling 0.06 0.04

0.4

0.02 0.00 -0.02 -16-14-12-10-8 -6 -4 -2 0 2 4 6 8

0.2

Crank Angle (deg)

0

-20

AHRR (kJ/deg)

1.2

Cylinder Pressure AHRR

AHRR (kJ/deg)

Pressure (bar)

120

0.0

-15

-10

-5

0

5

10

15

20

25

Crank Angle (deg)

E-50 Engine Performance and Exhaust Emissions Data Run # Fuel/Fuel % Main Fuel E Blend % Pilot Main SOI1 Dur us

g1_c2

g1_c3

g1_c4

2 E50/G50

3 E50/G50

4 E50/G50

1.00 0.50 0.00

1.00 0.50 0.00

1.00 0.50 0.00

950

950

950

25

AHRR (kJ/deg)

3 (bar) IMEPn Testing, E-50 Run2 100

239 Main SOI2 Dur us

0

0

0

Pilot Dur us

0

0

0

-15

-16

-17

-12.5

-13.5

-14.5

Dp/Dtheta

4.91

6.93

8.03

Fuel Rail Pressure [bar]

1000

1000

1000

brake emissions NOx_(g/kWhr)

5.09

4.67

5.94

14.42

14.75

16.40

19.51

19.42

22.34

0.00

0.00

0.01

CO_(g/kw-hr)

54.84

48.52

38.63

Intake_CO2_ (g/kW-hr)

42.17

36.55

34.90

Exhaust_CO 2_(g/kW-hr) SOF_(%)

1264 0.00

1149 0.00

1130 0.00

0.00

0.00

0.00

0.04

0.03

0.40

Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr)

AVL_Equival ence_Ratio AVL_C_(mg/ m^3)

240 emissions index emissions NOx_(g/kg-f)

9.76

9.85

13.13

HC_(g/kg-f) NOx_+_HC_( g/kg-f)

27.67

31.15

36.25

37.43

41.00

49.38

PM_(g/kg-f)

0.00

0.00

0.03

CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f]

105.23

102.44

85.38

2426

2427

2499

80.92

77.16

77.13

2.363

2.340

3.091

6.701

7.399

8.533

9.063

9.739

11.624

net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr)

0.001

0.001

0.007

25.479

24.331

20.098

Intake_CO2_ (g/ikW-hr)

19.592

18.327

18.157

Exhaust_CO 2_(g/ikW-hr)

587

576

588

Engine_Spee d_(RPM)

1300

1300

1300

Intake_Press ure_(psi)

30

30

30

Exhaust_Pre ssure_(psi)

31.4

31.3

31.4

241 Intake_Temp _(C)

100

100

100

Intake_Flowr ate_(kg/min)

2.967

2.967

2.967

70.93 0.85

71.14 0.68

71.15 0.66

4.25

4.25

4.25

469.71

404

375.95

20

22

23

3.69

4.06

4.25

0

0

0

0

0

7

521.16

473.66

452.49

20.25

22.27

23.28

1.40

1.54

1.61

1.92

1.92

1.92

Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow gasoline eq. [kg/hr] Fuel Flow [lbm/hr]

#DIV/0!

#DIV/0!

#DIV/0!

0.03

0.03

0.03

4.25

4.25

4.25

ETA thermal indicated

0.424

0.432

0.4362

ETA Gross Indicated

0.496

0.504

0.508

242

Carbon mass out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq)

1366

1417

1450.

13422

1340

1340

1.02

1.06

1.08

45.19

45.42

45.82

0.66

0.65

0.67

1.47

1.42

1.46

3.52

3.58

3.60

3.00

3.06

3.09

CA5 [ATDC]

5.78

4.55

3.59

CA10 [ATDC]

6.04

4.80

3.83

CA50 [ATDC]

7.75

6.12

5.11

CA75 [ATDC]

9.32

7.37

6.26

CA90 [ATDC]

11.58

9.57

8.37

CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur)

14.79

13.37

12.79

5.54

4.77

4.54

5.80

5.02

4.78

242.13

237.51

235.40

196.42

192.68

190.96

167.67

164.97

163.972

20.25

19.62

19.61

7.94 9.30

8.09 9.45

8.16 9.51

Standard_De viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar]

ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC Gas eq. gross EID [CA, SOI-CA50] Indicated Power [kw] Gross power

243 eta-comb (heywood 4.96)

0.0276

0.031

0.036

combustion efficiency

0.972

0.968

0.963

Carbon Based Equivalence Ratio phi A/F (A/F)s

0.127 92.57 11.77

0.127 92.57 11.77

0.127 92.57 11.77

244

6 (bar) IMEPn DTBP Tests 3.5% DTBP Indicated Cylinder Pressure and Apparent Heat Release Rate 3.5% DTBP 6 (bar) IMEPn, Run2

3.5% DTBP 6 (bar) IMEPn, Run3

1.0

0.015

100

0.010 0.005

0.8

0.000 -24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

60

0.6

40

0.4

20 0

0

5

10

15

20

0.010 0.005 -24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

60

0.6

40

0.4

0.2

20

0.2

0.0

0

25

0.0

-25 -20 -15 -10

3.5% DTBP 6 (bar) IMEPn, Run4

1.0

0.015

100

0.010

0.8

0.000 -24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

60

0.6

40

0.4

20 0

10

15

20

Pressure (bar)

0.005

5

80

0.010 0.005 -24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

40

0.4

0.2

20

0.2

0.0

0

0.0

-25 -20 -15 -10

5

10

15

20

25

100

0.010

0.8

0.000 -24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

60

0.6

40

0.4

20 0

15

20

Pressure (bar)

0.005

80

1.2 AHRR (kJ/deg)

1.0

0.015

Crank Angle (deg)

0

3.5% DTBP 6 (bar) IMEPn, Run7

AHRR (kJ/deg)

AHRR (kJ/deg)

Pressure (bar)

0.020

10

-5

120

Pressure AHRR

5

0.8

0.000

Crank Angle (deg)

NTC behavior

0

1.0

0.015

0.6

1.2

-5

25

Pressure AHRR

0.020

3.5% DTBP 6 (bar) IMEPn, Run6

-25 -20 -15 -10

20

60

25

120

80

15

NTC behavior

Crank Angle (deg)

100

10

1.2 AHRR (kJ/deg)

0.020

AHRR (kJ/deg)

AHRR (kJ/deg)

Pressure (bar)

NTC behavior

0

5

120

Pressure AHRR

-5

0

3.5% DTBP 6 (bar) IMEPn, Run5 1.2

-25 -20 -15 -10

-5

Crank Angle (deg)

120

80

0.8

0.000

Crank Angle (deg)

100

1.0

0.015

AHRR (kJ/deg)

-5

0.020

Pressure AHRR

NTC behavior 0.020

1.0

0.015 0.010 0.005

0.8

0.000 -24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

60

0.6

40

0.4

0.2

20

0.2

0.0

0

25

0.0

-25 -20 -15 -10

-5

0

5

10

Crank Angle (deg)

15

20

25

AHRR (kJ/deg)

-25 -20 -15 -10

80

1.2 Pressure AHRR

NTC behavior

AHRR (kJ/deg)

0.020

Pressure (bar)

80

NTC behavior

AHRR (kJ/deg)

Pressure (bar)

100

120

Pressure AHRR

AHRR (kJ/deg)

1.2 AHRR (kJ/deg)

120

245 3.5% DTBP 6 (bar) IMEPn, Run8

80

Pressure AHRR

NTC behavior 0.020

1.0

0.015 0.010 0.005

0.8

0.000 -24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

60

0.6

40

0.4

20

0.2

0

AHRR (kJ/deg)

Pressure (bar)

100

1.2 AHRR (kJ/deg)

120

0.0

-25 -20 -15 -10

-5

0

5

10

15

20

25

Crank Angle (deg)

3.5% 6 (bar) IMEPn DTBP Engine Performance and Exhaust Emissions Data g1_c2

g1_c3

g1_c4

g1_c5

g2_c2

g2_c3

g2_c5

2 E0_3.5 % DTBP 0.10 0.00 0.90

3 E0_3.5 % DTBP 0.10 0.00 0.90

4 E0_3.5 % DTBP 0.10 0.00 0.90

5 E0_3.5 % DTBP 0.10 0.00 0.90

6 E0_3.5 % DTBP 0.10 0.00 0.90

7 E0_3.5 % DTBP 0.10 0.00 0.90

8 E0_3.5 % DTBP 0.10 0.00 0.90

700

700

700

700

700

700

700

330 8000

330 8000

330 8000

330 8000

330 8000

330 8000

330 8000

-55

-55

-55

-55

-55

-55

-55

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

320

320

320

320

320

320

320

8.61

8.57

7.66

7.41

6.15

4.83

3.86

Fuel Rail Pressure [bar]

400.00

400.00

400.00

400.00

400.00

400.00

400.00

brake emissions Exhaust_Tem

199.00

199.00

199.00

199.00

199.00

199.00

199.00

Run #

Fuel/Fuel % Main Fuel E Blend % Pilot Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC] Dp/Dtheta

246 p_(C) NOx_(g/kWhr) HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr) CO_(g/kw-hr) Intake_CO2_ (g/kW-hr) Exhaust_CO 2_(g/kW-hr) SOF_(%) AVL_Equival ence_Ratio AVL_C_(mg/ m^3) emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_( g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f] net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr) Intake_CO2_ (g/ikW-hr) Exhaust_CO 2_(g/ikW-hr) Engine_Spee d_(RPM) Intake_Press

0.13 8.27

0.12 7.40

0.11 6.85

0.10 6.62

0.09 6.35

0.08 6.26

0.07 6.30

8.40

7.52

6.96

6.72

6.44

6.34

6.37

0.0090 16.90

0.0086 16.50

0.0072 16.29

0.0067 15.25

0.0066 14.18

0.0051 13.03

0.0036 11.90

10.26

58.90

89.08

125.34

152.96

192.30

211.02

616.59 0.00

609.92 0.00

600.73 0.00

609.26 0.00

613.99 0.00

619.39 0.00

602.93 0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.84

0.90

0.81

0.80

0.84

0.68

0.51

0.62 38.12

0.59 35.11

0.52 32.96

0.46 31.44

0.42 30.16

0.37 28.86

0.33 29.45

38.74 0.0416 77.92

35.70 0.0410 78.32

33.47 0.0345 78.38

31.90 0.0317 72.38

30.59 0.0315 67.38

29.22 0.0233 60.09

29.78 0.0166 55.66

2843

2895

2891

2892

2916

2855

2819

47.33

279.60

428.72

594.96

726.61

886.69

986.94

0.102

0.096

0.084

0.076

0.069

0.059

0.054

6.276

5.731

5.386

5.124

4.902

4.591

4.858

6.378

5.827

5.470

5.200

4.970

4.650

4.912

0.0069

0.0067

0.0056

0.0052

0.0051

0.0037

0.0027

12.829

12.784

12.808

11.798

10.948

9.561

9.182

7

45

70

96

118

141

162

468

472

472

471

473

454

465

1299 25.3

1300 25.3

1300 25.3

1300 25.3

1300 25.3

1300 25.3

1301 25.3

247 ure_(psi) Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr] ETA thermal indicated ETA Gross Indicated Carbon mass out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_De

26.6

26.7

26.7

26.6

26.6

26.6

26.6

40

40

40

40

40

40

40

2.912

2.687

2.344

2.44

2.2

2.022

1.896

79.19 0

73.05 8.44

68.2 13.76

63.67 19.64

59.69 24.09

54.82 30.34

51.38 34.37

40

40

40

40

40

40

40

0.62

0.62

0.62

0.62

0.62

0.62

0.62

21.62

20.29

21.55

22.43

21.69

21.34

22.59

68

70

71

70

70

68

69

12.56

12.93

13.11

12.93

12.94

12.56

12.74

5.4

5.4

5.4

5.4

5.4

5.4

5.4

0

0

7

7

7

7

7

216.88

210.67

207.78

210.67

210.51

216.88

213.81

68.84

70.87

71.88

70.87

70.87

68.84

69.85

4.75

4.89

4.96

4.89

4.89

4.75

4.82

2.72

2.72

2.72

2.72

2.72

2.72

2.72

0.05

0.05

0.05

0.05

0.05

0.05

0.05

6.02

6.02

6.02

6.02

6.02

6.02

6.02

0.506

0.510

0.510

0.511

0.513

0.524

0.505

0.545

0.550

0.549

0.550

0.552

0.563

0.544

2265

2337

2329

2306

2327

2270

2240

2351

2364

2372

2382

2390.3

2401

2406

0.96

0.99

0.98

0.97

0.97

0.95

0.93

92.82 2.64

93.61 1.654

93.51 2.82

93.74 2.68

94.02 2.85

3.13 3.36

92.62 3.66

248 viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur) ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC gross [g/kw-hr] Indicated Power [kw] Gross power [kw] eta-comb (heywood 4.96) combustion efficiency Carbon Based Equivalence Ratio phi A/F (A/F)s

2.85

2.51

3.01

2.86

3.03

3.36

3.95

6.74

6.80

6.79

6.81

6.83

6.96

6.73

6.26 -4.98 -3.73 0.30 2.03 4.27 5.52

6.31 -4.73 -3.48 0.55 2.29 4.75 6.25

6.31 -4.25 -3.00 1.02 2.55 4.76 5.77

6.32 -4.24 -2.75 1.32 3.28 6.01 7.76

6.34 -3.74 -2.24 2.06 4.04 6.52 8.25

6.48 -2.99 -1.48 3.30 5.52 8.26 10.00

6.25 -2.73 -0.99 4.26 6.53 9.27 11.01

8.00

8.24

7.76

8.76

8.76

9.74

10.26

9.25

9.49

9.01

10.24

10.26

11.25

12.00

164.63

163.22

163.40

162.99

162.49

159.12

164.97

164.63

163.22

163.40

162.99

162.49

159.12

164.97

52.80

53.05

53.52

53.82

54.56

55.80

56.76

16.55

16.69

16.67

16.71

16.76

17.12

16.51

17.82

17.98

17.96

18.00

18.05

18.40

17.79

0.056

0.053

0.051

0.048

0.045

0.042

0.042

0.9439

0.9468

0.9489

0.9518

0.9543

0.9573

0.9577

0.214 67.17 14.43

0.236 61.07 14.43

0.270 53.44 14.43

0.258 55.86 14.43

0.287 50.15 14.43

0.305 47.27 14.43

0.320 44.97 14.43

249

6 (bar) IMEPn DTBP Tests 1.75% DTBP Indicated Cylinder Pressure and Apparent Heat Release Rate 1.75% DTBP 6 (bar) IMEPn, Run2 160

100

1.6 Pressure AHRR 1.4

NTC behavior

0.020 0.015

1.2

0.010 0.005 0.000

1.0

-24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

20

0.2

0

AHRR (kJ/deg)

Pressure (bar)

120

AHRR (kJ/deg)

140

0.0

-25 -20 -15 -10

-5

0

5

10

15

20

25

Crank Angle (deg)

1.75% 6 (bar) IMEPn DTBP Engine Performance and Exhaust Emissions Data g1_c2 Run #

Fuel/Fuel % Main Fuel E Blend % Pilot Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

2 E0_1.7 5% DTBP 0.10 0.00 0.90 700 330 8000

-55 -36 -52.5 -33.5 320

250 Dp/Dtheta Fuel Rail Pressure [bar] brake emissions Exhaust_Tem p_(C) NOx_(g/kWhr) HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr) CO_(g/kw-hr) Intake_CO2_ (g/kW-hr) Exhaust_CO 2_(g/kW-hr) SOF_(%) AVL_Equival ence_Ratio AVL_C_(mg/ m^3) emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_( g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f] net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikw-

6.77

400.00

199.00 0.08 8.33 8.40 0.0075 20.09 10.10 617 0.00 0.00 0.67

0.34 37.43 37.77 0.0338 90.29 2777 45.40

0.056 6.064 6.119 0.0055 14.628

251 hr) Intake_CO2_ (g/ikW-hr) Exhaust_CO 2_(g/ikW-hr) Engine_Spee d_(RPM) Intake_Press ure_(psi) Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr] ETA thermal indicated ETA Gross Indicated Carbon mass

7.356 449

1300 25.3 26.7 40 2.908

79.09 0 32 0.61 24.06 65 12 5.29 0 222.47 65.80 4.54

2.67 0.04 5.90

0.514 0.554 2212

252 out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_De viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur) ISFC net [g/kw-hr] ISFC gross [g/kw-hr] ISFC gasoline eq. [g/kw-hr] Indicated Power [kw] Gross power [kw] eta-comb (heywood 4.96) combustion efficiency Carbon Based Equivalence Ratio phi A/F (A/F)s

2304 0.96

89.29

3.23 3.23 6.72 6.23 -4.73 -3.47 0.79 2.51 4.51 5.76 7.98 9.24

162.00 150.27

162.00 16.47 17.76

0.058 0.941

0.207 69.53 14.43

253

6 (bar) IMEPn DTBP Tests 0.75% DTBP Indicated Cylinder Pressure and Apparent Heat Release Rate 0.75% DTBP 6 (bar) IMEPn, Run2

80

Pressure AHRR

NTC behavior 0.020

1.0

0.015 0.010 0.005

0.8

0.000 -24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

60

0.6

40

0.4

20

0.2

0

AHRR (kJ/deg)

Pressure (bar)

100

1.2 AHRR (kJ/deg)

120

0.0

-25 -20 -15 -10

-5

0

5

10

15

20

25

Crank Angle (deg)

0.75% 6 (bar) IMEPn DTBP Engine Performance and Exhaust Emissions Data g1_c2 Run #

Fuel/Fuel % Main Fuel E Blend % Pilot Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC] Dp/Dtheta

2 E0_0.7 5% DTBP 0.10 0.00 0.90 700 330 8000

-55 -36 -52.5 -33.5 320 10.21

254

Fuel Rail Pressure [bar] brake emissions Exhaust_Tem p_(C) NOx_(g/kWhr) HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr) CO_(g/kw-hr) Intake_CO2_ (g/kW-hr) Exhaust_CO 2_(g/kW-hr) SOF_(%) AVL_Equival ence_Ratio AVL_C_(mg/ m^3) emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_( g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f] net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr)

400.00

199.00 0.10 7.75 7.85 0.0049 20.22 11.40 630 0.00 0.00 0.44

0.47 35.60 36.07 0.0225 92.91 2898 52.36

0.471 35.602 36.073 0.0225 92.914

255 Intake_CO2_ (g/ikW-hr) Exhaust_CO 2_(g/ikW-hr) Engine_Spee d_(RPM) Intake_Press ure_(psi) Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr] ETA thermal indicated ETA Gross Indicated Carbon mass out/hr

2898 52

1300 25.3 26.7 40 2.935

79.43 0 32 0.57 22.71 67 12.37 5.38 0 217.65 67.83 4.68

2.69 0.04 5.95

0.500 0.540

2318

256 carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_De viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur) ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC gross [g/kw-hr] Indicated Power [kw] Gross power [kw] eta-comb (heywood 4.96) combustion efficiency Carbon Based Equivalence Ratio phi A/F (A/F)s

2324 0.997

90.8

2.89 2.89 6.61 6.12 -3.99 -2.74 1.06 2.77 4.784 6.02 7.53 8.779

166.65

166.36 154.08 16.183 17.47

0.057 0.942

0.215 66.88 14.43

257

9 (bar) IMEPn DTBP Tests 3.5% DTBP Indicated Cylinder Pressure and Apparent Heat Release Rate 3.5% DTBP 9 (bar) IMEPn, Run2

120

0.020 0.015 0.010 0.005 0.000

1.0

-24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

100

0.8

60

0.6

40

0.4

20

0.2 0.0

0

0

5

10

15

20

25

0.005 0.000

80

0.8

60

0.6

40

0.4

20

0.2 0.0

-25 -20 -15 -10

140

Pressure AHRR 1.4

140

1.2

120

NTC behavior 0.015 0.010 0.005 0.000

1.0

-24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

20 0

0

5

10

15

20

Pressure (bar)

0.020

AHRR (kJ/deg)

Pressure (bar)

AHRR (kJ/deg)

160

-5

100

NTC behavior

0.005 0.000

0.6 0.4

0.2

20

0.2

0.0

0

25

0.0

-25 -20 -15 -10

-5

0

5

10

15

20

25

3.5% DTBP 9 (bar) IMEPn, Run7

1.2

120

1.0

-24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

20 0

15

20

100

AHRR (kJ/deg)

140

Pressure (bar)

0.000

Crank Angle (deg)

1.0

-24-22-20-18-16-14-12-10 -8 -6

40

0.005

10

1.2

0.010

60

160

0.010

5

1.6 Pressure AHRR 1.4

0.8

AHRR (kJ/deg)

AHRR (kJ/deg)

Pressure (bar)

NTC behavior

0

25

Crank Angle (deg)

1.6 Pressure AHRR 1.4

0.015

-5

20

80

0.020

-25 -20 -15 -10

15

0.015

3.5% DTBP 9 (bar) IMEPn, Run6

100

10

Crank Angle (deg)

160

120

5

0.020

Crank Angle (deg)

140

0

3.5% DTBP 9 (bar) IMEPn, Run5 1.6

AHRR (kJ/deg)

3.5% DTBP 9 (bar) IMEPn, Run4

-25 -20 -15 -10

-5

Crank Angle (deg)

160

100

1.0

-24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

120

1.2

0.010

AHRR (kJ/deg)

-5

0.015

1.6 Pressure AHRR 1.4

NTC behavior 0.020 0.015

1.2

0.010 0.005 0.000

1.0

-24-22-20-18-16-14-12-10 -8 -6

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

0.2

20

0.2

0.0

0

25

0.0

-25 -20 -15 -10

-5

0

5

10

Crank Angle (deg)

15

20

25

AHRR (kJ/deg)

-25 -20 -15 -10

NTC behavior 0.020

Crank Angle (deg)

80

0

1.6 Pressure AHRR 1.4

AHRR (kJ/deg)

1.2

140

Pressure (bar)

100

NTC behavior

160

AHRR (kJ/deg)

Pressure (bar)

120

AHRR (kJ/deg)

140

3.5% DTBP 9 (bar) IMEPn, Run3

1.6 Pressure AHRR 1.4

AHRR (kJ/deg)

160

258

3.5% 9 (bar) IMEPn DTBP Engine Performance and Exhaust Emissions Data g3_c2

g3_c3

g3_c4

g3_c5

g4_c2

g4_c3

2 E0_3.5 % DTBP 0.23 0.00 0.77

3 E0_3.5 % DTBP 0.19 0.00 0.81

4 E0_3.5 % DTBP 0.17 0.00 0.83

5 E0_3.5 % DTBP 0.14 0.00 0.86

6 E0_3.5 % DTBP 0.11 0.00 0.89

7 E0_3.5 % DTBP 0.10 0.00 0.90

1075

1040

990

940

800

775

560 9050

525 9250

500 9500

475 9750

430 10000

430 10150

-55

-55

-55

-55

-55

-55

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

320

320

320

320

320

320

Dp/Dtheta

11.20

7.92

7.91

6.40

6.22

4.60

Fuel Rail Pressure [bar]

400.00

400.00

400.00

400.00

400.00

400.00

199.00

199.00

199.00

199.00

199.00

199.00

0.07 3.32

0.06 3.35

0.06 3.26

0.06 3.26

0.05 3.40

0.05 3.79

3.38

3.41

3.32

3.32

3.45

3.83

0.0034 5.28

0.0033 5.50

0.0033 5.93

0.0033 5.93

0.0028 6.22

0.0024 6.59

314.39

312.80

313.20

311.57

305.22

309.63

629.75 0.00 0.00

629.00 0.00 0.00

630.13 0.00 0.00

630.13 0.00 0.00

615.09 0.00 0.00

622.72 0.00 0.00

Run #

Fuel/Fuel % Main Fuel E Blend % Pilot Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

brake emissions Exhaust_Tem p_(C) NOx_(g/kWhr) HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr) CO_(g/kw-hr) Intake_CO2_ (g/kW-hr) Exhaust_CO 2_(g/kW-hr) SOF_(%) AVL_Equival

259 ence_Ratio AVL_C_(mg/ m^3) emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_( g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f] net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr) Intake_CO2_ (g/ikW-hr) Exhaust_CO 2_(g/ikW-hr) Engine_Spee d_(RPM) Intake_Press ure_(psi) Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr)

1.00

0.97

0.95

0.95

0.82

0.72

0.35 16.97

0.33 17.57

0.31 17.10

0.31 17.06

0.27 17.95

0.25 19.61

17.29 0.0173 27.05

17.90 0.0172 28.90

17.42 0.0172 31.06

17.37 0.0171 30.98

18.22 0.0148 32.84

19.86 0.0125 34.13

3223

3303

3300

3292

3246

3226

1609

1642

1640

1628

1610

1604

0.057

0.054

0.050

0.049

0.044

0.040

2.776

2.811

2.706

2.694

2.877

3.159

2.828

2.864

2.756

2.744

2.921

3.199

0.0028

0.0028

0.0027

0.0027

0.0024

0.0020

4.425

4.623

4.915

4.893

5.264

5.496

263

262

259

257

258

258

527

528

522

520

520

519

1299

1300

1300

1300

1300

1300

25.3

25.3

25.3

25.3

25.3

25.3

26.7

26.7

26.7

26.7

26.6

26.6

38

38

38

38

38

38

1.411

1.419

1.417

1.413

1.41

1.414

55.04 49.69

55.37 49.5

55.33 49.47

55.33 49.47

54.96 49.39

55.15 49.49

40

40

40

40

40

40

1.89

1.55

1.35

1.15

0.92

0.8

260 BSFC_(g/kWhr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr] ETA thermal indicated ETA Gross Indicated Carbon mass out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_De viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC]

41.55

38.34

32.61

32.61

21.86

20.57

104

104

104

104

105

104

19.2

19.2

19.2

19.2

19.39

19.2

6.4

6.53

6.75

6.97

7.2

7.39

7

7

7

7

7

7

195.37

190.42

190.89

191.37

189.49

193.01

105.29

105.29

105.29

105.29

106.30

105.29

7.26

7.26

7.26

7.26

7.33

7.26

3.75

3.66

3.67

3.67

3.67

3.71

0.06

0.06

0.06

0.06

0.06

0.06

8.29

8.08

8.10

8.12

8.12

8.19

0.509

0.520

0.526

0.527

0.519

0.517

0.537

0.550

0.555

0.556

0.519

0.545

3404

3403

3411

3411

3370

3387

3319

3237

3245

3252

3251

3279

1.02

1.05

1.05

1.04

1.03

1.03

128.31

127.87

129.58

130.15

128.36

128.75

2.42

2.86

2.73

3.18

3.16

3.61

1.89

2.24

2.11

2.44

2.46

2.81

9.16

9.14

9.24

9.28

9.15

9.18

8.68 -1.99 -0.50 3.32 4.80 6.53 8.01

8.65 -0.96 0.76 5.06 6.81 9.00 10.52

8.76 -0.73 0.80 5.28 7.03 9.02 10.52

8.80 0.01 1.55 6.31 8.29 10.52 12.25

8.68 -0.50 1.26 6.07 8.05 10.31 12.01

8.71 -0.22 1.76 7.28 9.53 12.06 14.02

261 CA 10-90 (comb dur) CA 5-90 (comb dur) ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC gross [g/kw-hr] Indicated Power [kw] Gross power [kw] eta-comb (heywood 4.96) combustion efficiency Carbon Based Equivalence Ratio phi A/F (A/F)s

7.03

8.24

8.23

8.97

9.05

10.30

8.52

9.96

9.75

10.51

10.80

12.29

163.58

159.98

158.24

157.93

160.28

161.04

163.58

159.98

158.24

157.93

160.28

161.04

55.82

57.56

57.78

58.81

58.57

59.78

22.93

22.85

23.16

23.26

22.92

23.01

24.21

24.15

24.43

24.53

22.92

24.27

0.023

0.024

0.024

0.024

0.025

0.027

0.976

0.975

0.975

0.975

0.974

0.972

0.662 21.83 14.45

0.658 21.95 14.45

0.660 21.87 14.44

0.661 21.82 14.44

0.654 22.06 14.43

0.632 22.80 14.43

262

9 (bar) IMEPn DTBP Tests 1.75% DTBP Indicated Cylinder Pressure and Apparent Heat Release Rate 1.75% DTBP 9 (bar) IMEPn, Run2

120

0.010 0.005 0.000 -24-22-20-18-16-14-12-10 -8 -6

1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

20 0

-25 -20 -15 -10

-5

0

5

10

15

20

0.020 0.015 0.010 0.000 -24-22-20-18-16-14-12-10 -8 -6

100

0.8

60

0.6

40

0.4

0.2

20

0.2

0.0

0

25

0.0

-25 -20 -15 -10

0.020

160

1.2

120

140

0.010 0.005 0.000 -24-22-20-18-16-14-12-10 -8 -6

1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

20 0

-5

0

5

10

Crank Angle (deg)

15

20

Pressure (bar)

0.015

AHRR (kJ/deg)

AHRR (kJ/deg)

Pressure (bar)

NTC behavior

-25 -20 -15 -10

0

5

10

15

20

25

1.75% DTBP 9 (bar) IMEPn, Run5

1.6 Pressure AHRR 1.4

100

AHRR (kJ/deg)

1.75% DTBP 9 (bar) IMEPn, Run4

100

-5

Crank Angle (deg)

160

120

1.0

Crank Angle (deg)

80

Crank Angle (deg)

140

1.2

0.005

1.6 Pressure AHRR 1.4

NTC behavior 0.020 0.015 0.010

1.2

0.005 0.000 -24-22-20-18-16-14-12-10 -8 -6

1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

0.2

20

0.2

0.0

0

25

0.0

-25 -20 -15 -10

-5

0

5

10

Crank Angle (deg)

15

20

25

AHRR (kJ/deg)

100

1.6 Pressure AHRR 1.4

NTC behavior

AHRR (kJ/deg)

140

1.2

0.015

AHRR (kJ/deg)

0.020

Pressure AHRR 1.4

Pressure (bar)

Pressure (bar)

120

NTC behavior

160

AHRR (kJ/deg)

140

1.75% DTBP 9 (bar) IMEPn, Run3 1.6

AHRR (kJ/deg)

160

263 1.75% DTBP 9 (bar) IMEPn, Run6

1.75% DTBP 9 (bar) IMEPn, Run7

120

0.010 0.005 0.000

100

-24-22-20-18-16-14-12-10 -8 -6

1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

20 0

-25 -20 -15 -10

-5

0

5

10

15

20

1.6 Pressure AHRR 1.4

NTC behavior 0.020 0.015 0.010

1.2

0.005 0.000 -24-22-20-18-16-14-12-10 -8 -6

100

1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

0.2

20

0.2

0.0

0

25

AHRR (kJ/deg)

140

1.2

0.015

AHRR (kJ/deg)

0.020

Pressure AHRR 1.4

Pressure (bar)

Pressure (bar)

120

NTC behavior

160

AHRR (kJ/deg)

140

1.6 AHRR (kJ/deg)

160

0.0

-25 -20 -15 -10

Crank Angle (deg)

-5

0

5

10

15

20

25

Crank Angle (deg)

1.75% 9 (bar) IMEPn DTBP Engine Performance and Exhaust Emissions Data g1_c3

g1_c4

g1_c5

g2_c2

g2_c3

g2_c5

2 E0_1.7 5% DTBP 0.11 0.00 0.89

3 E0_1.7 5% DTBP 0.10 0.00 0.90

4 E0_1.7 5% DTBP 0.09 0.00 0.91

5 E0_1.7 5% DTBP 0.08 0.00 0.92

6 E0_1.7 5% DTBP 0.07 0.00 0.93

7 E0_1.7 5% DTBP 0.06 0.00 0.94

800

750

725

700

665

650

430 10000

380 10250

340 10350

310 10450

300 10500

290 10550

-55

-55

-55

-55

-55

-55

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

320

320

320

320

320

320

Dp/Dtheta

13.23

12.66

11.71

9.46

9.20

9.66

Fuel Rail Pressure [bar]

400.00

400.00

400.00

400.00

400.00

400.00

brake emissions Exhaust_Tem

199.00

199.00

199.00

199.00

199.00

199.00

Run #

Fuel/Fuel % Main Fuel E Blend % Pilot Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

264 p_(C) NOx_(g/kWhr) HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr) CO_(g/kw-hr) Intake_CO2_ (g/kW-hr) Exhaust_CO 2_(g/kW-hr) SOF_(%) AVL_Equival ence_Ratio AVL_C_(mg/ m^3) emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_( g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f]

0.05 3.31

0.05 3.34

0.04 3.31

0.04 3.41

0.04 3.48

0.04 3.57

3.37

3.38

3.35

3.45

3.52

3.61

0.0051 6.41

0.0044 5.71

0.0069 5.65

0.0077 5.46

0.0062 6.41

0.0038 7.53

276.85

269.09

259.73

265.15

261.18

260.18

631.60 0.00

618.73 0.00

605.95 0.00

619.37 0.00

609.18 0.00

599.49 0.00

0.00

0.00

0.00

0.00

0.00

0.00

1.24

1.12

1.75

1.92

1.55

0.96

0.28 17.27

0.24 18.00

0.23 17.81

0.22 17.99

0.21 18.39

0.21 19.01

17.56 0.0266 33.41

18.25 0.0238 30.78

18.04 0.0370 30.41

18.21 0.0406 28.85

18.60 0.0328 33.81

19.21 0.0200 40.04

3293

3259

3189

1397

3270 1400.0 0

3214

1443

3335 1450.7 2

1378

1384

0.044

0.038

0.036

0.035

0.033

0.034

2.648

2.781

2.801

2.840

2.913

3.037

2.692

2.818

2.838

2.875

2.946

3.070

0.0041

0.0037

0.0058

0.0064

0.0052

0.0032

net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr) Intake_CO2_ (g/ikW-hr) Exhaust_CO 2_(g/ikW-hr)

5.123

4.755

4.783

4.556

5.356

6.397

221 504.97 1

224 515.21 2

219 512.61 4

221 516.34 2

218 509.27 9

221 509.49 0

Engine_Spee d_(RPM) Intake_Press

1300 25.3

1300 25.3

1300 25.3

1300 25.3

1300 25.3

1300 25.3

265 ure_(psi) Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr] ETA thermal indicated ETA Gross Indicated Carbon mass out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_De

26.7

26.7

26.6

26.6

26.6

26.5

32

32

32

32

32

32

1.652

1.669

1.683

1.678

1.678

1.678

64.55 43.52

65.18 43.18

65.91 42.55

65.74 42.5

65.77 42.56

65.38 43.08

32

32

40

40

40

40

0.9

0.79

0.69

0.65

0.54

0.5

22.55

17.63

16.18

15.13

12.55

9.73

102

107

106

105

105

106

18.85

19.76

19.57

19.4

19.39

19.57

7.09

7.31

7.35

7.47

7.58

7.63

0

7

7

7

7

7

191.80

185.48

185.90

189.39

189.49

187.98

103.26

108.32

107.31

106.30

106.30

107.31

7.12

7.47

7.40

7.33

7.33

7.40

3.62

3.67

3.64

3.67

3.67

3.68

0.06

0.06

0.06

0.06

0.06

0.06

7.99

8.10

8.04

8.12

8.12

8.13

0.543

0.539

0.529

0.527

0.525

0.521

0.573

0.569

0.557

0.557

0.555

0.550

3351

3438

3333

3388

3342

3332

3192

3233

3207

3240

3239

3242

1.05

1.06

1.04

1.04

1.03

1.03

128.55 1.635

129.44 2.07

129.41 2.87

130.17 2.9

129.84 2.86

128.81 4.82

266 viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur) ISFC net [g/kw-hr] ISFC gross [g/kw-hr] ISFC gasoline eq. [g/kw-hr] Indicated Power [kw] Gross power [kw] eta-comb (heywood 4.96) combustion efficiency Carbon Based Equivalence Ratio phi A/F (A/F)s

2.01

2.07

2.22

2.23

2.86

3.74

9.41

9.46

9.23

9.29

9.26

9.17

8.91 -1.95 -0.49 3.08 4.37 5.840 6.836

8.97 -1.22 0.26 4.07 5.52 7.067 8.055

8.75 -0.50 1.02 5.10 6.61 8.368 9.531

8.80 -0.23 1.32 5.55 7.08 8.768 9.80

8.78 0.08 1.80 6.25 7.8 9.35 10.35

8.71 1.05 3.04 8.61 10.80 13.00 14.31

6.33

6.80

7.34

7.44

7.55

9.97

7.78

8.28

8.86

9.00

9.28

11.91

153.34

154.45

157.26

157.88

158.41

159.75

145.37

146.48

149.16

149.58

150.16

151.76

153.34

154.45

157.26

157.88

159.75

23.57

23.73

23.13

23.27

158.41 23.193 47

24.86

25.02

24.38

24.56

24.46

24.236

0.0249

0.0251

0.0248

0.0246

0.0262

0.0282

0.9750

0.9748

0.9751

0.9753

0.9737

0.9717

0.557 25.89 14.43

0.565 25.50 14.43

0.543 26.52 14.42

0.552 26.12 14.42

0.544 26.50 14.42

0.542 26.59 14.42

23.02

267

9 (bar) IMEPn DTBP Tests 0.75% DTBP Indicated Cylinder Pressure and Apparent Heat Release Rate

0.020

1.2

120

140

0.010 0.005 0.000 -24-22-20-18-16-14-12-10 -8 -6

1.0

Crank Angle (deg)

0.8

60

0.6

40

0.4

20

0.2 0.0

0

-25 -20 -15 -10

-5

0

5

10

15

20

0.015 0.010 0.000 -24-22-20-18-16-14-12-10 -8 -6

80

0.8

60

0.6

40

0.4

20

0.2

25

0.0

-25 -20 -15 -10

0.020

160

1.2

120

140

0.010 0.005 0.000 -24-22-20-18-16-14-12-10 -8 -6

1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

20 0

-5

0

5

10

15

20

Pressure (bar)

0.015

AHRR (kJ/deg)

AHRR (kJ/deg)

Pressure (bar)

NTC behavior

-25 -20 -15 -10

0

5

10

15

20

25

0.75% DTBP 9 (bar) IMEPn, Run5

1.6 Pressure AHRR 1.4

AHRR (kJ/deg)

0.75% DTBP 9 (bar) IMEPn, Run4

100

-5

Crank Angle (deg)

160

120

1.0

Crank Angle (deg)

Crank Angle (deg)

140

1.2

0.005

100

80

0

0.020

1.6 Pressure AHRR 1.4

NTC behavior 0.020 0.015 0.010

1.2

0.005 0.000

100

-24-22-20-18-16-14-12-10 -8 -6

1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

0.2

20

0.2

0.0

0

25

0.0

-25 -20 -15 -10

Crank Angle (deg)

AHRR (kJ/deg)

100

Pressure (bar)

0.015

1.6 Pressure AHRR 1.4

NTC behavior

AHRR (kJ/deg)

NTC behavior

160

AHRR (kJ/deg)

Pressure (bar)

120

AHRR (kJ/deg)

140

0.75% DTBP 9 (bar) IMEPn, Run3

1.6 Pressure AHRR 1.4

AHRR (kJ/deg)

0.75% DTBP 9 (bar) IMEPn, Run2 160

-5

0

5

10

15

20

25

Crank Angle (deg)

0.75% 9 (bar) IMEPn DTBP Engine Performance and Exhaust Emissions Data Run #

Fuel/Fuel % Main Fuel E Blend % Pilot

g1_c3

g1_c4

g1_c5

g2_c2

2 E0_0.7 5% DTBP 0.11 0.00 0.89

3 E0_0.7 5% DTBP 0.10 0.00 0.90

4 E0_0.7 5% DTBP 0.08 0.00 0.92

5 E0_0.7 5% DTBP 0.06 0.00 0.94

268 Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us

800

750

700

640

430 10000

380 10250

310 10450

310 10550

-55

-55

-55

-55

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

-36 -52.5 -33.5

320

320

320

320

Dp/Dtheta

26.16

16.95

11.38

6.67

Fuel Rail Pressure [bar]

400.00

400.00

400.00

400.00

199.00

199.00

199.00

199.00

0.07 3.15

0.06 3.31

0.04 3.36

0.04 3.51

3.22

3.37

3.40

3.55

0.0028 6.07

0.0035 6.14

0.0031 7.20

0.0023 7.83

279

272

263

263

639 0.00

634 0.00

616 0.00

616 0.00

0.00

0.00

0.00

0.00

0.71

0.86

0.77

0.57

0.36 16.53

0.30 17.25

0.22 17.55

0.21 18.26

16.89 0.0149 31.90 3356

17.55 0.0184 32.01 3304

17.78 0.0160 37.66 3224

18.47 0.0120 40.68 3200

Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

brake emissions Exhaust_Tem p_(C) NOx_(g/kWhr) HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr) CO_(g/kw-hr) Intake_CO2_ (g/kW-hr) Exhaust_CO 2_(g/kW-hr) SOF_(%) AVL_Equival ence_Ratio AVL_C_(mg/ m^3) emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_( g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2

269 [g/kg-f] INT CO2 [g/kg-f] net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr) Intake_CO2_ (g/ikW-hr) Exhaust_CO 2_(g/ikW-hr) Engine_Spee d_(RPM) Intake_Press ure_(psi) Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc

1464

1419

1377

1367

0.361

0.298

0.225

0.212

16.531

17.254

17.554

18.258

16.893

17.551

17.780

18.470

0.0149

0.0184

0.0160

0.0120

31.903

32.013

37.663

40.677

3356

3304

3224

3200

1464

1419

1377

1367

1299

1300

1300

1300

25.3

25.3

25.3

25.3

26.7

26.7

26.6

26.7

32

32

32

32

1.658

1.69

1.69

1.685

63.98 43.31

65.41 42.63

65.74 42.4

65.64 42.41

32

32

40

40

0.88

0.81

0.64

0.5

22.88

19

14.1

10.92

103

105

104

103

19.02

19.39

19.2

19.02

7.12

7.41

7.47

7.59

0

7

7

7

190.32

191.82

191.13

192.46

104.27 7.19

106.30 7.33

105.29 7.26

104.27 7.19

270 [Bar] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr] ETA thermal indicated ETA Gross Indicated Carbon mass out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_De viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur) ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC gross [g/kw-hr] Indicated Power [kw] Gross power

3.62

3.72

3.67

3.66

0.06

0.06

0.06

0.06

8.00

8.22

8.11

8.09

0.527

0.533

0.523

0.518

0.556

0.563

0.552

0.547

3413

3456

3338

3325

3197

3281

3236

3228

1.07

1.05

1.03

1.03

128.07

129.94

128.89

127.31

1.63

1.88

2.87

4.00

2.61

1.88

2.23

3.14

9.15

9.51

9.20

9.09

8.66 -1.22 0.26 3.54 4.79 6.26 7.04

9.01 -0.49 1.01 4.58 6.03 7.54 8.51

8.72 0.52 2.04 6.30 8.03 10.01 11.29

8.61 1.03 3.00 8.03 10.02 12.05 13.51

6.01

6.53

7.97

9.04

7.49

8.02

9.50

11.02

158.12

156.17

159.28

160.87

158.12

156.17

159.28

160.87

149.69

148.04

150.88

152.35

22.89 24.18

23.82 25.13

23.04 24.32

22.75 24.03

271 [kw] eta-comb (heywood 4.96) combustion efficiency Carbon Based Equivalence Ratio phi A/F (A/F)s

0.02

0.02

0.03

0.03

0.98

0.98

0.97

0.97

0.57 25.50 14.43

0.56 25.69 14.43

0.54 26.61 14.43

0.54 26.75 14.42

272

9 (bar) IMEPn DTBP Tests E-98 Diesel Indicated Cylinder Pressure and Apparent Heat Release Rate E-98 9 (bar) IMEPn, Run2

E-98 9 (bar) IMEPn, Run3 140

1.2 -15

-10

1.0

-5

Crank Angle (deg)

120 100

0.8

60

0.6

40

0.4

20

0.2 0.0

0

-25 -20 -15 -10

-5

0

5

10

15

20

0.020 0.016 0.012 0.008 0.004 0.000 -0.004 -20

1.2 -15

-10

0.8

60

0.6

40

0.4

20

0.2

25

0.0

-25 -20 -15 -10

E-98 9 (bar) IMEPn, Run4 160

1.4

140

1.2 -10

1.0

-5

Crank Angle (deg)

120

Pressure (bar)

-15

100

0.8

60

0.6

40

0.4

20

0.2 0.0

0

0

-5

0

5

10

10

15

20

25

15

20

Pressure AHRR

NTC behavior 0.020 0.016 0.012 0.008 0.004 0.000 -0.004 -20

1.4 1.2

-15

-10

1.0

-5

Crank Angle (deg)

80

-25 -20 -15 -10

5

1.6

AHRR (kJ/deg)

1.6

AHRR (kJ/deg)

100

AHRR (kJ/deg)

Pressure (bar)

120

0

E-98 9 (bar) IMEPn, Run5 Pressure AHRR

NTC behavior

-5

Crank Angle (deg)

160

0.020 0.016 0.012 0.008 0.004 0.000 -0.004 -20

1.0

-5

80

Crank Angle (deg)

140

1.4

Crank Angle (deg)

80

0

1.6 Pressure AHRR

NTC behavior

80

0.8

60

0.6

40

0.4

20

0.2

25

0.0

-25 -20 -15 -10

Crank Angle (deg)

-5

0

5

10

15

20

25

Crank Angle (deg)

E-98 Diesel 9 (bar) IMEPn Engine Performance and Exhaust Emissions Data g1_c2 Run # Fuel/Fuel % Main Fuel % Pilot Main Eneregy (kJ/hr) Pilot Eneregy (kJ/hr)

AHRR (kJ/deg)

1.4

g1_c3

g1_c4

g1_c5

2

3

4

5

E-98/D 0.26 0.74 128.43 234.58

E-98/D 0.25 0.75 123.76 238.40

E-98/D 0.24 0.76 116.95 246.04

E-98/D 0.23 0.77 113.98 248.22

AHRR (kJ/deg)

0.020 0.016 0.012 0.008 0.004 0.000 -0.004 -20

160

Pressure (bar)

100

AHRR (kJ/deg)

Pressure (bar)

120

NTC behavior

1.6

AHRR (kJ/deg)

140

Pressure AHRR

AHRR (kJ/deg)

160

273 Main fuel energy percent Pilot fuel energy percent Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us

0.35 0.65 730 495 11250

0.34 0.66 850 350 11000

0.32 0.68 875 375 10500

0.31 0.69 875 375 10500

Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

-30 -28 -53 -37 -320

-35 -25 -58 -34 -320

-35 -25 -58 -34 -320

-35 -25 -58 -34 -320

Dp/Dtheta

4.55

5.30

5.71

6.54

800.00

800.00

800.00

800.00

-4.00 0.34

-4.00 0.30

-4.00 0.24

-4.00 0.25

Fuel Rail Pressure [bar] brake emissions Exhaust_Temp_(C) NOx_(g/kW-hr) HC_(g/kW-hr) NOx_+_HC_(g/kW-hr) PM_(g/kW-hr) CO_(g/kw-hr) Intake_CO2_(g/kW-hr) Exhaust_CO2_(g/kW-hr) SOF_(%) AVL_Equivalence_Ratio AVL_C_(mg/m^3) emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_(g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f] net indicated emissions NOx_(g/ikW-hr) HC_(g/ikW-hr) NOx_+_HC_(g/ikW-hr) PM_(g/ikW-hr) CO_(g/ikw-hr) Intake_CO2_(g/ikW-hr) Exhaust_CO2_(g/ikW-hr) Engine_Speed_(RPM)

N/A

N/A

3.06 0.01 12.82 7.03 597.36 0.00 0.00 1.28

N/A

8.89 0.01 11.95 5.98 590.88 0.00 0.00 1.30

1.25

N/A

6.81 0.01 12.18 6.31 596.51 0.00 0.00 1.00

-3.65 0.01 9.16 6.60 599.04 0.00 0.00 1.58

0.87

11.27 0.03 47.26 2202.92 25.93

32.97 0.03 44.33 2191.64 22.19

24.59 0.03 44.00 2154.81 22.79

0.90 N/A -13.20 0.04 33.16 2167.71 23.88

0.286 2.285 2.571 0.008 10.787 5.917 502.749

0.251 7.095 7.346 0.008 9.877 4.945 488.316

0.199 5.411 5.610 0.006 10.039 5.199 491.660

0.206 -3.208 -3.002 0.009 7.542 5.431 493.012

1299

1300

1300

1300

N/A

1.13 N/A

N/A

274 Intake_Pressure_(psi) Exhaust_Pressure_(psi) Intake_Temp_(C) Intake_Flowrate_(kg/min) Upstream_Orifice_Press_(psi) EGR_(%) EGR_Temp_(C) HP_Fuel_Flowrate_(lb/hr) BSFC_(g/kW-hr) Brake_Torque_(ft-lbs) Brake_Power_(kW) LP_Fuel_Flowrate_(lb/hr) VFD_Setpoint(Hz) BSFC_(g/kW-hr) BMEP calc [Psi] BMEP calc [Bar]

25.3 27.1 32 2.95 81.03 0 0 11.62 3.02 3.02 70.55 105 19.39 8.6 5.363334 0

25.3 27.2 36 2.933 79.93 0 0 11.64 2.91 2.7 68.66 104 19.2 8.74 5.450644 0

25.3 27.1 40 2.969 79.03 0 0 11.77 2.75 2.5 66.3 102 18.83 9.02 5.625264 0

25.3 27.2 44 2.933 77.96 0 0 11.78 2.68 2.4 64.45 102 18.83 9.1 5.675156 0

Fuel Flow gasoline equlivent [kg/hr] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr]

3.793364 5.26 0.09 11.62

3.783097 5.18 0.09 11.44

3.789712 5.21 0.09 11.52

3.780613 5.20 0.09 11.50

0.505 0.534

0.519 0.548

0.510 0.536

0.512 0.537

ETA thermal indicated ETA Gross Indicated Carbon mass out/hr carbon mass in/hr Carbon balance

3318.3 3234.7 1.03

3357.4 3174.1 1.06

3285.4 3167.2 1.04

3077.1 3151.2 0.98

IMEP_(psi) (endaq) Standard_Deviation_of_IMEP(psi) COV_of_IMEP(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur)

128.89 4.17 3.24 9.22 8.72 -4.91 -2.74 4.30 6.81 9.00 10.28 11.75 16.66

129.88 3.85 2.97 9.30 8.78 -4.45 -2.38 4.55 6.81 8.81 10.07 11.19 15.65

127.93 3.7 2.89 9.18 8.65 -3.69 -1.48 5.04 7.12 9.13 10.53 10.60 14.30

127.93 3.64 2.85 9.17 8.65 -3.17 -0.99 5.06 7.05 8.86 10.07 9.84 13.02

ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC gross gas eq. [g/kw-hr] Indicated Power [kw] Gross power [kw]

164.65 228.22 155.58 23.04 24.38

162.83 222.81 153.79 23.23 24.60

165.88 228.17 156.29 22.85 24.25

165.24 227.43 155.81 22.88 24.26

275

9 (bar) IMEPn DTBP Tests E-98 Diesel Indicated Cylinder Pressure and Apparent Heat Release Rate 160

1.4

140

1.2

Crank Angle (deg)

120

Pressure (bar)

1.0

-20 -18 -16 -14 -12 -10 -8 -6

100

0.8

60

0.6

40

0.4

20

0.2 0.0

0

0

5

10

15

20

25

0.8

60

0.6

40

0.4

20

0.2 0.0

-25 -20 -15 -10

1.4

140

1.2

120

1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

20 0

-25 -20 -15 -10

-5

0

5

10

15

20

100

0.4

0.2

20

0.2

0.0

0

25

0.0

-25 -20 -15 -10

160

1.4

140

1.2

120

0.8

60

0.6

40

0.4

20 0

Crank Angle (deg)

-5

0

5

10

15

20

15

20

100

Pressure AHRR AHRR (kJ/deg)

1.6

1.0

10

1.0

0.6

Crank Angle (deg)

5

1.2

40

80

0

1.4

60

Pressure (bar)

NTC behavior

-5

NTC behavior 0.020 0.016 0.012 0.008 0.004 0.000 -20 -18 -16 -14 -12 -10 -8 -6

1.6

25

E-98 9 (bar) IMEPn, Run7

0.020 0.016 0.012 0.008 0.004 0.000 -20 -18 -16 -14 -12 -10 -8 -6

-25 -20 -15 -10

25

0.8

AHRR (kJ/deg)

100

20

Crank Angle (deg)

Pressure AHRR AHRR (kJ/deg)

Pressure (bar)

120

15

Crank Angle (deg)

E-98 9 (bar) IMEPn, Run6 140

10

80

Crank Angle (deg)

160

5

Pressure AHRR AHRR (kJ/deg)

160

Pressure (bar)

NTC behavior 0.020 0.016 0.012 0.008 0.004 0.000 -20 -18 -16 -14 -12 -10 -8 -6

1.6

AHRR (kJ/deg)

100

0

E-98 9 (bar) IMEPn, Run5 Pressure AHRR

AHRR (kJ/deg)

Pressure (bar)

120

-5

Crank Angle (deg)

E-98 9 (bar) IMEPn, Run4 140

1.0

80

Crank Angle (deg)

160

1.2

-20 -18 -16 -14 -12 -10 -8 -6

AHRR (kJ/deg)

-5

1.4

NTC behavior 0.020 0.016 0.012 0.008 0.004 0.000 -20 -18 -16 -14 -12 -10 -8 -6

1.6 1.4 1.2 1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

0.2

20

0.2

0.0

0

25

0.0

-25 -20 -15 -10

-5

0

5

10

Crank Angle (deg)

15

20

25

AHRR (kJ/deg)

-25 -20 -15 -10

NTC behavior 0.010 0.008 0.006 0.004 0.002 0.000

1.6

Crank Angle (deg)

80

0

Pressure AHRR AHRR (kJ/deg)

NTC behavior 0.010 0.008 0.006 0.004 0.002 0.000

1.6

AHRR (kJ/deg)

Pressure (bar)

100

AHRR (kJ/deg)

140 120

E-98 9 (bar) IMEPn, Run3 Pressure AHRR

AHRR (kJ/deg)

E-98 9 (bar) IMEPn, Run2 160

276 E-98 9 (bar) IMEPn, Run9 1.4

140

1.2

120

1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

20 0

-25 -20 -15 -10

-5

0

5

10

15

20

100

Pressure AHRR AHRR (kJ/deg)

100

NTC behavior 0.020 0.016 0.012 0.008 0.004 0.000 -20 -18 -16 -14 -12 -10 -8 -6

160

Pressure (bar)

Pressure (bar)

120

AHRR (kJ/deg)

140

1.6

AHRR (kJ/deg)

Pressure AHRR

NTC behavior 0.020 0.016 0.012 0.008 0.004 0.000 -20 -18 -16 -14 -12 -10 -8 -6

1.6 1.4 1.2 1.0

Crank Angle (deg)

80

0.8

60

0.6

40

0.4

0.2

20

0.2

0.0

0

25

AHRR (kJ/deg)

E-98 9 (bar) IMEPn, Run8 160

0.0

-25 -20 -15 -10

Crank Angle (deg)

-5

0

5

10

15

20

25

Crank Angle (deg)

E-98 Diesel 9 (bar) IMEPn Engine Performance and Exhaust Emissions Data Run # Fuel/Fuel % Main Fuel % Pilot Main Eneregy (kJ/hr) Pilot Eneregy (kJ/hr) Main fuel energy percent Pilot fuel energy percent Main SOI1 Dur us Main SOI2 Dur us Pilot Dur us

g1_c2

g1_c3

g1_c4

g1_c5

g2_c2

g2_c3

g2_c4

g3_c1

1 E-98/D 0.24 0.76

2 E-98/D 0.27 0.73

3 E-98/D 0.30 0.70

4 E-98/D 0.30 0.70

5 E-98/D 0.30 0.70

6 E-98/D 0.30 0.70

7 E-98/D 0.30 0.70

8 E-98/D 0.30 0.70

121.63

135.24

149.70

149.70

149.70

149.70

149.70

149.70

237.99

229.90

217.77

217.77

217.77

217.77

217.77

217.77

0.34

0.37

0.41

0.41

0.41

0.41

0.41

0.41

0.66

0.63

0.59

0.59

0.59

0.59

0.59

0.59

850

850

875

875

875

875

875

875

350 11500

350 11000

375 10500

375 10500

375 10500

375 10500

375 10500

375 10500

-30

-35

-35

-35

-35

-35

-35

-35

Main SOI1 [ATDC] Main SOI2 [ATDC] Actual SOI1 Actual SOI2 Pilot SOI [ATDC]

-28 -53 -37

-25 -58 -34

-25 -58 -34

-25 -58 -34

-25 -58 -34

-25 -58 -34

-25 -58 -34

-25 -58 -34

-320

-320

-320

-320

-320

-320

-320

-320

Dp/Dtheta

4.60

4.44

10.00

10.00

9.13

8.02

4.90

3.94

277 Fuel Rail Pressure [bar] brake emissions Exhaust_Tem p_(C) NOx_(g/kWhr) HC_(g/kW-hr) NOx_+_HC_( g/kW-hr) PM_(g/kWhr) CO_(g/kw-hr) Intake_CO2_ (g/kW-hr) Exhaust_CO 2_(g/kW-hr) SOF_(%) AVL_Equival ence_Ratio AVL_C_(mg/ m^3) emissions index emissions NOx_(g/kg-f) HC_(g/kg-f) NOx_+_HC_( g/kg-f) PM_(g/kg-f) CO_(g/kg-f) Exh CO2 [g/kg-f] INT CO2 [g/kg-f] net indicated emissions NOx_(g/ikWhr) HC_(g/ikWhr) NOx_+_HC_( g/ikW-hr) PM_(g/ikWhr) CO_(g/ikwhr) Intake_CO2_

800.00

800.00

800.00

800.00

800.00

800.00

800.00

245.00

248.00

249.00

250.00

250.00

250.00

0.58 N/A

0.26 N/A

0.69 N/A

0.77 N/A

0.51 N/A

0.42 N/A

0.26 N/A

0.20 N/A

2.16

1.76

2.13

2.15

1.78

1.72

2.12

2.05

0.01 11.50

0.02 13.80

0.03 7.59

0.03 6.58

0.02 6.89

0.02 7.38

0.01 9.93

0.01 10.11

7.34

7.49

7.20

42.67

96.31

122.72

150.87

172.34

611.83 0.00

605.27 0.00

614.39 0.00

617.42 0.00

605.16 0.00

612.77 0.00

605.32 0.00

603.04 0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

1.78

2.31

4.62

4.15

3.58

3.14

1.65

1.43

2.06 N/A

0.90 N/A

2.54 N/A

2.80 N/A

1.90 N/A

1.54 N/A

0.95 N/A

0.71 N/A

7.68 0.05 40.92 2178.0 1

6.20 0.06 48.59 2131.5 6

7.80 0.12 27.76 2247.3 8

7.88 0.10 24.06 2258.4 7

6.62 0.08 25.70 2256.2 8

6.34 0.07 27.26 2263.6 6

7.75 0.03 36.31 2213.0 6

7.36 0.03 36.26 2163.3 8

26.13

26.39

26.35

156.07

359.10

453.34

551.60

618.26

0.472

0.205

0.565

0.623

0.422

0.339

0.213

0.162

1.283

1.208

1.170

1.127

1.048

1.063

1.521

1.511

1.754

1.412

1.735

1.749

1.470

1.402

1.734

1.673

0.0105

0.0137

0.0268

0.0226

0.0176

0.0146

0.0073

0.0061

9.346 5.967

11.059 6.006

6.171 5.858

5.342 34.648

5.706 79.734

6.025 100.18

8.128 123.47

8.241 140.52

250-4

800.00

248.00

278 (g/ikW-hr) Exhaust_CO 2_(g/ikW-hr) Engine_Spee d_(RPM) Intake_Press ure_(psi) Exhaust_Pre ssure_(psi) Intake_Temp _(C) Intake_Flowr ate_(kg/min) Upstream_Or ifice_Press_( psi) EGR_(%) EGR_Temp_( C) HP_Fuel_Flo wrate_(lb/hr) BSFC_(g/kWhr) Brake_Torqu e_(ft-lbs) Brake_Power _(kW) LP_Fuel_Flo wrate_(lb/hr) VFD_Setpoin t(Hz) BSFC_(g/kWhr) BMEP calc [Psi] BMEP calc [Bar] Fuel Flow gasoline equlivent [kg/hr] Fuel Flow [kg/hr] Fuel Flow [kg/min] Fuel Flow [lbm/hr] ETA thermal indicated ETA Gross Indicated

497.39 0

485.15 8

499.61 1

501.37 8

500.98 4

7 500.26 8

0 495.37 4

4 491.71 8

1299

1300

1299

1301

1300

1300

1299

1300

25.3

25.3

25.3

25.3

25.3

25.3

25.3

25.3

27.3

27.2

27.4

27.3

27.3

27.3

27.1

27.3

32

32

32

32

32

32

32

32

2.97

2.93

2.897

2.858

2.858

2.91

2.64

2.33

80.72 0

81.09 0

80.51 0

75.84 6.04

68.46 15.21

64.96 19.39

60.96 24.35

58.01 28.05

32

32

32

32

32

32

32

32

2.86

3.18

3.52

3.52

3.52

3.52

3.52

3.52

67.56

77.25

82.61

83.77

81.14

82.86

82.78

84.69

102

101

104

104

106

105

104

102

18.83

18.66

19.2

19.2

19.57

19.39

19.19

18.83

8.83

8.53

8.08

8.08

8.08

8.08

8.08

8.08

0

0

0

0

0

0

0

0

280.91

283.96

273.38

273.38

268.21

270.70

273.52

278.75

103.26

102.25

105.29

105.29

107.31

106.30

105.29

103.26

7.12

7.05

7.26

7.26

7.40

7.33

7.26

7.12

3.7858 7

3.8460 09

3.8728 68

3.8728 68

3.8728 68

3.8728 68

3.8728 68

3.8728 68

5.29

5.30

5.25

5.25

5.25

5.25

5.25

5.25

0.09

0.09

0.09

0.09

0.09

0.09

0.09

0.09

11.69

11.71

11.60

11.60

11.60

11.60

11.60

11.60

0.508

0.503

0.508

0.508

0.508

0.511

0.504

0.496

0.538

0.541

0.554

0.554

0.554

0.557

0.549

0.542

279

Carbon mass out/hr carbon mass in/hr Carbon balance IMEP_(psi) (endaq) Standard_De viation_of_IM EP(psi) COV_of_IME P(pct) Gross IMEP [bar] IMEP NET [bar] CA5 [ATDC] CA10 [ATDC] CA50 [ATDC] CA75 [ATDC] CA90 [ATDC] CA95 [ATDC] CA 10-90 (comb dur) CA 5-90 (comb dur) ISFC net [g/kw-hr] ISFC gasoline eq. [g/kw-hr] ISFC gross gas eq. [g/kw-hr] Indicated Power [kw] Gross power [kw]

3260.2

3214.7

3303.6

3310.2

3309.0

3323.4

3280.0

3208.1

3199.5

3253.3

3279.3

3288.9

3303.6

3310.8

3318.5

3324.3

1.02

0.99

1.01

1.01

1.00

1.00

0.99

0.97

127.93

128.57

130.39

130.57

130.55

131.05

129.6

127.54

3.86

3.77

3.65

3.94

3.59

3.89

3.8

3.91

3.02

2.93

2.8

3.02

2.75

2.97

2.94

3.07

9.27

9.33

9.47

9.48

9.47

9.51

9.40

9.27

8.76 -3.24 -1.49 5.55 8.27 11.63 15.76

8.81 -2.92 -1.15 5.77 8.53 12.12 16.04

8.93 -5.19 -3.69 1.85 3.84 7.38 13.77

8.94 -5.47 -4.13 1.36 3.32 6.85 13.53

8.94 -5.15 -3.65 2.05 4.06 7.76 14.01

8.98 -4.69 -3.17 2.58 4.76 8.34 14.27

8.88 -3.20 -1.49 5.02 7.76 11.52 15.63

8.74 -2.48 -0.72 6.31 9.51 13.37 17.10

13.12

13.28

11.07

10.98

11.41

11.51

13.01

14.09

16.37

16.20

16.26

16.45

16.56

16.20

16.21

16.56

228.37

227.61

222.31

222.00

222.04

221.00

223.84

227.29

163.45

165.21

164.03

163.80

163.83

163.06

165.16

167.71

154.43

155.88

154.67

154.50

154.64

153.88

155.92

158.03

23.16

23.28

23.61

23.64

23.64

23.75

23.45

23.09

24.52

24.67

25.04

25.07

25.04

25.17

24.84

24.51

Suggest Documents