Improving Engine Performance by Optimizing Fuel ...

THIESEL 2010 Conference on Thermo- and Fluid Dynamic Processes in Diesel Engines

Improving Engine Performance by Optimizing Fuel Reactivity with a Dual Fuel PCCI Strategy 1

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D.A. Splitter , R.M. Hanson , S.L. Kokjohn and R.D. Reitz 1

Engine Research Center, University of Wisconsin-Madison, 1500 Engineering Dr., Madison, WI 53706, USA. E-mail: Telephone: Fax:

[email protected] +(1) 608 263 1596 +(1) 608 263 9870

Abstract. In this work engine experiments and multi-dimensional modelling were used to explore a dual-fuel concept to realize highly-efficient premixed charge compression ignition (PCCI) combustion with near zero levels of NOx and soot. In-cylinder fuel blending using port-fuel-injection of gasoline and optimized direct-injection of diesel fuel was used to control combustion phasing and duration. In addition to injection and operating parameters, the study explored the effect of fuel properties by considering both gasoline-diesel dual fuel operation, and ethanol (E85)-diesel dual fuel operation. Remarkably, high gross indicated thermal efficiencies were achieved, reaching 59% and 56% for E85diesel and gasoline-diesel, respectively. Multi-dimensional modelling, using the KIVA-CHEMKIN code, was used to select the operating conditions for the dual-fuel engine experiments over a range of loads from near idle to 16.5 bar IMEPg. The modelling results showed that as the engine load increased, the fuel reactivity requirements to achieve optimum combustion phasing decreased. Engine experiments were then performed on a heavy-duty test engine to validate the model predictions and the modelling was further used to explain the experimentally observed results. The experiments confirmed that by optimizing the fuel reactivity based on the specific operating conditions, combustion phasing can be optimized in order to minimize fuel consumption. Additionally, it was found that highly efficient operation (greater than 50% indicated thermal efficiency) could be achieved with both gasoline-diesel and ethanol-diesel dual fuel blends over a wide range of loads. This study showed that, compared to gasoline-diesel, significantly higher quantities of diesel fuel were required to maintain optimal combustion phasing with the ethanoldiesel fuel blends. This result is due to a combination of the lower reactivity and higher enthalpy of vaporization of ethanol (compared to gasoline) and combustion chemistry effects of ethanol diesel blends.

1. Introduction As an energy conversion device the internal combustion engine has become invaluable to society. The ability of the internal combustion engine to provide economically viable and reliable power for both stationary and mobile applications has resulted in mass production of the internal combustion engine in a variety of displacement and design platforms. However, with such widespread use, engine related emissions and fuel consumption have become of societal, political, and technical concern. Specifically research and development to simultaneously reduce green house gas emissions and fuel consumption has come to the forefront. Although several solutions have been researched, a majority of the potential solutions can be grouped into what have become known as low temperature combustion (LTC) strategies. LTC strategies have been demonstrated by a variety of researchers, with a selection in Refs. [[1-7],[12],[8] [40]]. To maintain high thermal efficiency the described LTC strategies are compression ignited (CI) and operate un-throttled with diesel-like compression ratios. However, unlike traditional diesel combustion, LTC strategies can also reduce engine out emissions through the use of long ignition delay, thereby lowering local equivalence ratios and temperatures. Decreased local equivalence ratios and temperatures, reduce the formation of particulates (soot) and oxides of nitrogen (NOx) [[8][9] Although some of the above references [[1-7],[8],[12][40]] have demonstrated several successfully proven LTC strategies, these strategies have been historically either limited to mid and low-

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D. A. Splitter, R. M. Hanson, S. L. Kokjohn, R. D. Reitz

loads due to high combustion noise or technically challenging high load control [[11], [40]]. Although low and medium load operation is more common in light duty applications, heavy duty applications operate and are emissions-regulated at high engine loads. The need to operate at such loads with efficient clean combustion has presented itself as a challenge for LTC strategies. In the present research a strategy for mid and high load LTC operation is explored through engine experiments and detailed CFD modeling. In order to avoid high pressure rise rate (PRR) and extend the premixed combustion duration, in-cylinder fuel blending of different fuels is used as a charge preparation strategy. The strategy consists of direct-injection of a more reactive fuel (diesel), and port fuel injection of a less reactive fuel (gasoline or gasoline ethanol blends). Through this charge preparation strategy, operation with greater than 50% thermal efficiency while meeting EPA 2010 heavy duty emissions mandates has previously been demonstrated in [13] and [14] with acceptable PRR. Unlike similar dual fuel compression ignition demonstrated by Inakagki et al. [15], the optimized blending strategy in the present research has been shown to result in a staged combustion process [16], thus significantly extending the premixed combustion duration. In this work, the previous studies (i.e., [13] and [14]) are expanded upon to explore high load operation and fuel effects.

2. Experimental design

2.1 Engine Laboratory Design All the experiments were conducted using a heavy-duty 2.44 L Caterpillar 3401 Single Cylinder Oil Test Engine (SCOTE). The engine geometry is summarized in Table 1 and the lab setup is shown in Fig.1.

Figure 1 Diagram of the engine lab setup

Improving Engine Performance by Optimizing Fuel Reactivity with a Dual Fuel PCCI Strategy

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Table 1 Caterpillar 3401E SCOTE engine geometry Displacement (L) Geometric Compression Ratio (-) Effective Compression Ratio (-) Squish Height (mm) Bowl Volume (mL) Connecting Rod Length (mm) Bore x Stroke (mm) Piston Swirl Ratio Valvetrain

2.44 16.1:1 15.1-13.7:1 1.97 110.8 261.6 137.2 x 165.1 Standard SCOTE (open crater) Stock = 0.7 EVC = -355 deg (° CA ATDC) compression IVC = -143 to -85 deg (° CA ATDC) EVO =130 deg (° CA ATDC) compression IVO =335 deg (° CA ATDC) compression

To lower peak cylinder pressures and aid in combustion phasing control, different methods of modifying intake valve closing (IVC) timing were investigated by Nevin [17], who used four different custom manufactured camshafts with IVC timings ranging from -143º ATDC (stock) to -85º ATDC in order to lower the effective compression ratio. The intake valve lift profiles are shown in Fig. 2. Note that the effective compression ratio of the -85º ATDC cam is by definition is approximately 9:1 (by geometry). However, it was found to provide an effective compression ratio of 13.7:1 (calculated by a GT-POWER motoring simulation), a result of the feathered lift profile seen in Fig. 2. For high load operation, KIVA-CFD modeling results suggested the nominal IVC timing of -85º ATDC would be optimal, due to the resulting lower TDC temperatures and pressures. Thus, the -85º ATDC cam was used exclusively for all loads above a nominal load of 9 bar IMEP. The engine experiments were performed using port fuel injection of gasoline and multiple early direct injections of diesel fuel with a conventional (i.e., wide angle, large nozzle hole) Bosch high-pressure, common-rail injector. The specifications for the port fuel injector and common rail injector are given in Tables 2 and 3, respectively. 16 14

Lift [mm]

12 10

85 IVC85

8 6

143 IVC143

4 2 0 -25

25

75

125

175

225

275

Crank Angle [ATDC] IVC143

IVC130

IVC115

IVC100

IVC85

Figure 2 Intake valve lift profiles for custom camshafts

Table 2 Port fuel injector specifications Injection Pressure (MPa) Included Spray Angle (°) Number of Holes Steady Flowrate (cc/min)

0.5 15 3 750

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Table 3 Common rail injector specifications Injection Pressure (MPa) Included Spray Angle (°) Number of Holes Hole Diameter (µm) Steady Flowrate (cc/30 sec) at 100 bar K-factor (-)

80 145 6 250 1000 0

Cylinder pressure was measured with a Kistler model 6067C1 water-cooled pressure transducer in conjunction with a Kistler model 510 charge amplifier. Acquired cylinder pressure traces were averaged for 500 cycles, and low-pass filtered at 2200 Hz to aid in smoothing pressure data. The intake air flow rate was measured using choked flow orifices. To obtain a choked flow condition for a variety of engine operating conditions, combinations of 6 different sized orifices were used to gain the desired intake air flow rate. Intake air was heated with two immersion style heaters and used PID to control the temperature to within ±1°C. Both the in take and exhaust system pressures were equipped with PID control to maintain the pressures ±0.7 kPa. PM measurements were performed with an AVL model 415s smoke meter. PM measurements of filter smoke number (FSN), concentration (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. All gaseous emissions measurements were performed with a conventional 5-gas emissions bench. The EGR rate was determined through the ratio of intake CO2 to exhaust CO2 levels. Gaseous emissions were averaged for 30 seconds after attaining a steady-state operating condition for several minutes. The EGR system consisted of an electrically driven supercharger and a diesel particulate filter (DPF), to prevent potential fouling of the EGR cooler and supercharger. The EGR supercharger was implemented as a pump to maintain constant EGR levels with constant surge tank pressures as the DPF fills. Conversely, the use of the supercharger in this manner allows more repeatable and stable EGR rates. Also, the EGR supercharger allows any EGR rate to be achieved, but it was used such that realistic intake and exhaust conditions were simulated for the delivered EGR rate. 2.2 Fuel Properties Commercially available 91 pump octane number (PON†) gasoline and ultra low sulfur diesel (ULSD) were used for all tests. The ethanol fuel consisted of 98% anhydrous ethanol and 2% natural gasoline denaturant by volume, obtained directly from the distiller. Independent lab analysis of the diesel and gasoline fuel properties is seen in Tables 4 and 5 respectively. Ethanol fuel blends in the tests are referred as EX, with X being the percent ethanol by volume, i.e., E85 is 85% ethanol and 15% gasoline by volume. Table 4 Diesel fuel properties Specific Gravity (@ 15.5 °C) (-) Viscosity (@ 40 °C) (cSt) Surface Tension (@ 25 °C) (dyne/cm) Lower Heating Value (MJ/kg) Cetane Number H/C ratio

†

PON=(RON+MON)/2, also referred to as the anti knock index

0.856 2.71 30 42.526 46.1 1.74


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Table 5 Gasoline fuel properties Distillation Curve Initial Boiling Point (°C) Temperature 10% evaporated (°C) Temperature 50% evaporated (°C) Temperature 90% evaporated (°C) Final Boiling Point (°C) Lower Heating Value (MJ/kg) MON RON (RON+MON)/2 Ethanol (%) H/C ratio Specific Gravity (@ 15.6 °C) (-) Sulfur (ASTM D5453) (PPM)

ASTM D86 38.89 69.4 105 160.56 215.56 43.22 87.8 95.6 91.6 0 1.88 0.737 4.6

Table 6 Ethanol fuel properties† Ethanol Lower Heating Value (MJ/kg) Enthalpy of Vapoorization (kJ/kg) Ethanol Octane Number (RON) H/C ratio (-) E85 is 85% Ethanol and 15% Gasoline

26.9 840 107 3.0 By vol.

2.3 Operating Conditions Engine experiments of dual-fuel engine operation were performed using direct injection of a more reactive fuel (diesel) and port fuel injection of less reactive fuel (gasoline or gasoline ethanol blend). A similar strategy has been demonstrated to be successful by Inagaki et al. [15], at low and mid engine loads. However, unlike their work for the present study the fuel delivery strategy used was selected through the KIVA-CFD optimization of Kokjohn et al. [13]. This study demonstrated that for the SCOTE engine geometry, the port and direct-injection timings depicted in Table 7 were optimal. These timings were experimentally verified to be optimal and used both in the previous studies of [13][14][16] as well as in the present study. Furthermore, the use of multiple injections was shown to dramatically increase engine and combustion efficiency. Table 7 Optimized Charge Preparation Strategy Port fueling percentage Port injection pressure (bar) Direct injection pressure (bar) SOI1 fuel percentage (%) SOI2 fuel percentage (%) Port Injection Timing (° ATDC) SOI1 (first direct injection) Injection Timing (° ATDC) SOI2 (second direct injection) Injection Timing (° ATDC)

†

Properties as defined in Table D.4 of [37]

varied with load 4.14 800 60 40 -360 -55 -36

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4. Computational

Simulation

Computations were performed using the KIVA-3v release 2 code [18] with improvements to many physical and chemistry models developed at the ERC [19][20][21]. The KIVA-3v code is coupled with the CHEMKIN II solver for detailed chemistry calculations. A 49 species and 179 reaction mechanism describing the oxidation of iso-octane, n-heptane, and ethanol [22] was used to simulate gasoline, diesel fuel, and ethanol chemistry. Many studies (e.g., Ra et al. [23]) have shown the combustion characteristics of gasoline and diesel are represented well by iso-octane (i.e., PRF 100) and n-heptane (i.e., PRF 0), respectively. This approach has also been shown to yield acceptable agreement for blends of gasoline and diesel fuel (e.g., Kokjohn et al.[13]). The physical properties (for spray and mixing processes) of diesel fuel are represented by tetradecane. This modeling approach has been shown to yield acceptable results in numerous studies (e.g., Kong et al. [24]). Soot is predicted using a phenomenological soot model [24] based on the approach of Hiroyasu [25]. The soot model used in the present study uses acetylene as an inception species, which allows the soot model to be coupled to the chemistry solver through the addition of 13 reactions involving acetylene. NOx emissions are predicted using a reduced NO mechanism [26] consisting of 4 additional species and 12 reactions. The reduced NO mechanism is based on the Gas Research Institute (GRI) NO mechanism [27]. The spray model employed in this study uses the Lagrangian-Drop and Eulerian-Fluid (LDEF) approach. Because a detailed chemistry model is used, it is desirable to use a relatively coarse computational mesh; however, severe grid size dependency has been observed in LDEF spray models. This problem is most severe in the near nozzle region where the droplets are very close together and occupy only a small portion of the Eulerian mesh. Abraham [28] showed that accurate modeling of the near nozzle region required grid resolution on the order of the orifice diameter. However, it is not feasible from a computational time standpoint to solve engine problems on such a fine mesh. In order to reduce the grid size dependency of the LDEF spray model and allow accurate spray simulation on a relatively coarse grid, the Gasjet model of Abani et al. [19] is employed to model the relative velocity between the droplets and gas phase in the near nozzle region. Droplet breakup is modeled using the hybrid Kelvin Helmholtz (KH) – Rayleigh Taylor (RT) model described by Beale and Reitz [29]. The droplet collision model is based on O’Rourke’s model; however, a radius of influence method is used to determine the possible collision partners to further reduce mesh dependency [19]. In addition, the collision model was expanded by Munnannur [30] to include a more comprehensive range of collision outcomes. The current implementation of the droplet collision model considers bounce, coalescence, and fragmenting and non-fragmenting separations. Droplet interactions with the wall are considered through a wall film submodel [31][32], which includes the effects associated with splash, film spreading, and motion due to inertia. A three dimensional computational grid with the crevice volume resolved, as shown in Fig. 3 was used for all simulations. To reduce computational burden, a 60 degree sector mesh (corresponding to a single nozzle hole of the six hole injector of this study) with periodic boundaries was used.

Figure 3. Computational grid with crevice volume shown at 20° BTDC. The grid has 18,000 cells at BDC.


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4. Results and Discussion As discussed in the operating conditions section, engine experiments of dual-fuel engine operation were performed using the optimized timings given in Table 7. Although Hanson et al. [14] demonstrated engine loads of 9 and 11 bar IMEPn with gasoline port and diesel direct injection fueling, investigations at both higher engine loads and fuel effects were performed. Initially the investigations of higher engine load with gasoline port and diesel direct injection were conducted. 4.1 Mid and High Load Gasoline-Diesel Operation Operating the engine with the injection strategy described in Table 7 at 1300 rev/min, the total fueling was progressively increased to increase engine load. The tested loads can be seen in Table 8 where at each operating condition exhaust emissions were sampled, and indicated cylinder pressure measurements were acquired. Select indicated cylinder pressure and apparent heat release rate (AHRR) traces of this load sweep can be seen in Fig. 4. While changing load, the 50% mass fraction burned combustion timing (CA 50) was allowed to adjust as needed as to retain PRR below approximately 10 bar/deg, and maximum power. Although usually a lower pressure rise rate was demonstrated, a 10 bar/deg maximum PRR was selected because it is representative of typical heavy-duty engine operation. To more quantitatively examine the performance of the combustion strategy throughout the load sweep, combustion and emissions trends are plotted as a function of load in Fig. 5, where it can be seen that the peak PRR remained low for all tested loads. Also, emissions remained low throughout the sweep, with all mid-and low-load cases passing EPA 2010 HD emissions standards in cylinder. Although high load operation nearly passed soot mandates, Fig. 5 shows that the equivalence ratio at these higher load conditions is at unity (as also indicated by the sudden turn-up in CO emissions), demonstrating extremely low soot emissions even at stoichiometric CI operation. Furthermore, Fig. 5 demonstrates extremely low fuel consumption while operating with low emissions, lower PRR and low COV†. The indicated specific fuel consumption trends of Fig. 5 demonstrate a local minimum, of approximately 150 g/kW-hr (gross based), at the 6 to 12 bar IMEPg operating conditions. The low fuel consumption at these loads are a result of more extensive optimization performed at these loads [[12][13][14]]. It is believed that further optimization of the higher load operation points could yield further fuel consumption gains. Another operating condition that was not optimized for performance was the lowest load operating condition of 5.2 bar IMEPg. This operating condition is demonstrated in [16], where optical tests of the combustion process were conducted. In that research optics were installed into a modified cylinder head that required extremely low PRR (~4 bar/deg) and peak cylinder pressure for structural considerations. In that case the combustion timing was adjusted from the optimal efficiency timing to maintain the PRR. Table 8 Gasoline-diesel Respective Load Operating Conditions Engine gross IMEP (bar) 5.2 6.7 9.6 11.6 13.6 Total fueling (kg/hr) 2.34 2.72 3.97 4.73 6.00 Port fuel percentage mass and energy (%) 68 89 78 87 83 EGR rate (%) 0 0 37 37 57 Absolute intake pressure (kPa) 137.8 174.3 174.3 191.6 219.1 Absolute exhaust pressure (kPa) 144.7 184.7 184.7 201.9 235.7 Camshaft IVC timing (°ADTC) -143 -143 -85 -85 -85

†

COV as defined by the standard deviation of the IMEP by the mean of the IEMP

14.5 6.26 90 57 234.3 251.5 -85


5.2 bar IMEPg, 0% EGR | 11.6 bar IMEPg, 37% EGR | 14.2 bar IMEPg, 43% EGR 160

Pressure (bar)

120

1.5

NTC Behavior

AHRR (kJ/° CA)

140

9.6 bar IMEPg, 37% EGR 13.4 bar IMEPg, 43% EGR

100

0.05 0.04 0.03 0.02 0.01 0.00 -20 -16 -12 -8 -4 Crank Angle (° CA ATDC)

1.3 1.1

AHRR (kJ/° CA)

8

0.9

80

0.8

60

0.6

40

0.4 Increasing Load

20

0.2

0

-0.1

-25 -20 -15 -10

-5

0

5

10

15

20

25

Crank Angle (° CA ATDC)

PM (g/kW-hr)

20 18 16 14 12 10 8 6

0.015

10 6

4.0 3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0

0.3 0.2

2010 HD limit 0.1 0.0 4

6

8

10

12

14

IMEPg (bar)

16

18

COV (%)

2

12 10 8 6 4 2 0 -2

180

0.60 0.58 0.56 0.54 0.52 0.50 0.48 0.46

60 50 40 30 20 10 0

8 7 6 5 4 3 2 1 0

16

170

12

160

8

150

4

140

Φ (−)

0.000

η g indicated (-)

NOx (g/kW-hr)

PRR (bar/° CA)

0.005 14

CA50 (° CA ATDC)

0.010

gasoline/diesel

0

1.0

EGR (%)

gasoline/diesel

0.020

2010 HD limit

CO (g/kw-hr)

gasoline/diesel

ISFCg gasoline eq.(g/kW-hr) HTHR 5-90 (° CA) HC (g/kw-hr)

Figure 4 Gasoline-diesel dual fuel operation at various loads. Note that the low temperature heat release events are similar in magnitude, with slightly later phasing at higher loads due to additional charge cooling provided by greater port injected fuel mass. Also, intake pressure is increased at the 13.4 and 14.2 bar IMEPg loads as to retain equivalence ratio of unity.

0.8 0.6 0.4 0.2 0.0

4

6

8

10

12

14

IMEPg (bar)

16

18

4

6

8

10

12

14

16

18

IMEPg (bar)

Figure 5 Emissions and combustion trends of gasoline-diesel load sweep. Note the high thermal efficiency with low emission, PRR, and COV, demonstrating the benefits of tailoring fuel reactivity over a wide range of loads. The demonstrated gasoline-diesel operation has shown that fuel reactivity controlled PCCI is a successful operating strategy for high engine efficiency and low emissions from light to high engine loads. Although engine loads of 14.5 bar IMEPg were demonstrated with gasoline-diesel operation, the maximum load of the engine was not realized due to limitations of the lab EGR system. As seen in Fig. 5, the peak cylinder pressure at 14.5 bar IMEPg operation with a gasoline-diesel fueling strategy was 140 bar, while peak maximum continuous cylinder pressure of the test engine is rated at 175 bar. Although peak load was not obtained, there were no combustion related phenomena that prevented operation at higher engine loads. The combustion duration has been shown through KIVA-CFD simulations ([13] and [14]), and measured spectra [16] to be controlled by the reactivity gradient, such that the fuel decomposition process occurs first in areas of high local fuel reactivity and progresses to areas of less reactive fuel.


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The present study has demonstrated this phenomenon with a reactivity gradient generated by 91.6 PON pump gasoline and #2 ULSD diesel fuel. Although these fuels are common, they are only a sample of commercially common and available fuels. Exploring this combustion process with different fuels with various reactivities and chemistry is desired to more fully explore reactivity controlled combustion. 4.2 Mid and High Load E85-Diesel Operation

4.2.1 Ethanol Operating Considerations and Combustion Processes To expand on the successful mid-and high-load results with gasoline-diesel fueling, testing with gasoline-ethanol blends was selected for two reasons. Primarily, it was desired to examine the effect of altering the reactivity gradient produced by the dual-fuel charge preparation strategy, and ethanol is a high octane fuel [37]. Secondly, ethanol is often added to pump gasoline, where it is commonly available in blends of ethanol with gasoline in ratios as high as 85% by volume (E85). To explore the effect of a greater reactivity gradient, port-fuel-injection of E85 was used in place of gasoline port-fuelinjection, but diesel direct-injection was maintained. The direct injection timings were un-altered from those previously demonstrated with gasoline-diesel operation and shown in Table 7. Although these direct injection timings were optimized with KIVA-CFD for gasoline-diesel operation, they were experimentally validated to be correctly timed at mid-load for E85-diesel operation as well. However, unlike the injection timings, port fueling of E85-diesel required some alterations to the engine operating conditions account for fuel property differences. To account for heating value differences of E85 compared to gasoline, fueling was adjusted accordingly to meet the load requirements. Also, the intake temperature was increased from 32°C to 42°C to help account for the engines response to en thalpy of vaporization differences between the two fuels. The corresponding engine conditions for E85-diesel operation are seen in Table 9. It can be seen that similar intake and exhaust pressures were used with E85-diesel and gasoline-diesel operation. For a given load, the effects of lower energy density, charge temperature, and fuel reactivity of E85 as compared to gasoline required significantly more diesel fuel energy percentage. Research by Hashimoto [34] has demonstrated that ethanol n-heptane mixtures exhibit significantly different low temperature chemistry. In that research measurements of natural luminosity of low temperature heat release in RCM combustion with mixtures of 30% ethanol by volume and n-heptane were shown to exhibit significantly lower luminosity and corresponding pressure during low temperature reaction, suggesting chemical pathway differences. The combination of these four effects required approximately twice the diesel fuel percentage (by energy) for a given load with port fueling of E85 as compared to port fueling of gasoline. Figure 6 demonstrates this requirement as a function of load by plotting both energy percent and mass percent of both E85 and gasoline port-fueling strategies. Also noted in Fig. 6 is that as engine load increases, the in-cylinder blending of fuel reactivity becomes overall less reactive. Tailoring the fuel reactivity in this manner, in combination with EGR, reduces PRR at higher engine loads. From Fig. 6, the higher energy percentage of direct-injected diesel fuel with E85-diesel would suggest that the low temperature reactions (low temperature heat release) would be more prevalent with E85 blends. This trend was seen in previous dual-fuel results [14] where at a constant load, additional gasoline (and thus lower diesel fuel percentage) showed to lower the LTHR and increase the HTHR. However, with E85-diesel operation this trend was not seen. For example, at the engine load of 9.6 bar IMEPg, Fig. 7 demonstrates that the low temperature heat release was significantly decreased. This is a similar trend that was shown by Hashimoto [34]. In that research it was suggested through CHEMKIN simulation using the n-heptane mechanism of Curran et al. [36], and the ethanol mechanism of Marinov [35], that ethanol inhibited low temperature reaction chemistry. As seen in Fig. 7, the present research demonstrates that E85 operation exhibits similar suppressed low temperature heat release, but also results in faster transition and longer duration of high temperature heat release. This combustion process is significantly different from that of gasoline-diesel operation. Although lengthened combustion duration is also observed in the propagation of premixed flames, it is believed at this load that the lengthened combustion duration of E85 operation is not a result of flame propagation. To support flame propagation the global equivalence ratio must be high enough. Ethanol has been shown by Warnatz et al. [41], Egolfopoulus et al. [38], and Röhl et al. [39], to have similar flame speeds and slightly richer flammability limits as diesel and gasoline like fuels. At the 9.6 bar IMEPg condition, the global equivalence ratio is low (approximately 0.3), and would not allow significant flame

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structures or burning velocities to support flame propagation. To gain further insight into the possible differences of E85 combustion as compared to gasoline, multi-dimensional simulations with KIVA-CFD coupled with the CHEMKIN solver were performed, as shown in the next section. Table 9 E85-Diesel Load Operating Conditions Engine gross IMEP (bar) Total fueling (kg/hr) Port fuel mass percentage (%) Port fuel energy percentage (%) EGR rate (%) Absolute intake pressure (kPa) Absolute exhaust pressure (kPa) Camshaft IVC timing (°ADTC)

9.6 4.81 79 72 0 201.9 212.2 -85

11.6 6.16 82 75 0 204.7 215.0 -85

14.9 8.52 82 75 44 220.5 231.5 -85

15.5 8.98 83 77 43 224.0 234.3 -85

16.5 9.57 83 77 47 246.0 256.4 -85

E-85/Diesel Port Fuel Energy Percent E-85/Diesel Port Fuel Mass Percent Gasoline/Diesel Port Fuel Energy and Mass Percent

100

Port Fuel Percent (%)

95 90 85 80 75 70 65 60 4

6

8

10

12

14

16

18

IMEPg (bar)

Figure 6 Port fuel energy and mass percentages for gasoline-diesel, and E85-diesel operation. Note that the mass and energy percentage of gasoline-diesel are a similar due to the nearly identical heating values between the fuels.

4.2.2 Mid-Load Multi-Dimensional Simulation Results with Ethanol To understand the differences between dual-fuel PCCI combustion using gasoline-diesel fuel blends and E85-diesel fuel blends, detailed CFD modeling was performed. The operating conditions for gasoline-diesel and E85-diesel, at a nominal 9 bar IMEPg load, are shown in Table 10. However, the simulated gasoline-diesel case differs from that shown in Table 8. A different gasoline-diesel case was selected so as to better match the trapped mass of the E85-diesel case. The better matched case was taken from Hanson et al. [14] using identical fuel delivery strategy to that of Table 7, but operated with the differences presented in Table 10. Although these two cases are different, the differences were found to not result in any significant differences in emissions, efficiency or combustion. Figure 7 shows a comparison of the measured and predicted cylinder pressure and apparent heat release rates for operation at the nominal 9 bar IMEPg and 1300 rev/min using both E85-diesel and gasoline-diesel fuel blends. It can be seen that the simulations do an excellent job capturing the combustion characteristics for both blends. More specifically, notice that the change in combustion characteristics between gasoline-diesel and E85-diesel is captured well by the simulations. For instance, dual-fuel PCCI combustion using a blend of E85 and diesel fuel shows a nearly “triangular shaped” heat release profile and a very broad combustion duration, quite different from the more symmetric gasoline-diesel result.


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Table 10 Operating conditions of nominal 9 bar IMEPg gasoline-diesel case of [14] (presented in Fig. 7). Operating case

Gasoline-diesel Condition in Table 8 9.6 3.97 78 78 37 174.3 184.7 -85 0.00359

Engine gross IMEP (bar) Total fueling (kg/hr) Port fuel mass percentage (%) Port fuel energy percentage (%) EGR rate (%) Absolute intake pressure (kPa) Absolute exhaust pressure (kPa) Camshaft IVC timing (°ADTC) Trapped mass (kg)

Pressure [MPa]

12

Gasoline-diesel Condition in [14] 9.2 3.67 89 89 43 174.3 184.0 -143 0.00496

E85-diesel Condition in Table 9 9.6 4.81 79 72 0 201.9 212.2 -85 0.00500

E85 and Diesel Fuel

10

Gasoline and Diesel Fuel

8 6

E85 & Diesel - Experiment E85 & Diesel - Simulation Gasoline & Diesel - Experiment Gasoline & Diesel - Simulation

4 2 0

AHRR [J/deg]

600 500 400 300

AHRR [J/deg]

700 50 40 30 20 10 0

LTHR

-14 -12 -10

-8

Crank [° CA ATDC]

200 100 0 -20 -15 -10

-5

0

5

10

15

20

Crank [° CA ATDC]

Figure 7 Comparison of measured and predicted cylinder pressure and apparent heat release rate (AHRR). The solid lines are experiments and dashed lines are modeling predictions. Note the differences in low and high temperature heat release rates with E85 as compared to gasoline port fueling, even with twice as much diesel fuel used with E85. To understand the differences in the combustion characteristics between E85-diesel fuel blends and gasoline-diesel fuel blends, Fig. 8 shows the evolutions of several key species for both fueling strategies. It can be seen that, in general, the combustion processes between gasoline-diesel fuel and E85diesel fuel blends are very similar. Specifically, n-heptane (diesel surrogate in the computations) is consumed early in the cycle resulting in a significant buildup in formaldehyde, which was also experimentally validated in [16]. As the combustion processes transition to second stage ignition, a simultaneous consumption in formaldehyde, isooctane, and, in the case of E85-diesel fuel blends, ethanol is observed. However, several differences can be found. It appears that the consumption of n-heptane is delayed for the ethanol-diesel fuel blends compared to the gasoline-diesel fuel blends. The delayed consumption of n-heptane for the E85-diesel fuel case results in a delay in the buildup of formaldehyde. Additionally, the transition to second stage ignition (signaled by the consumption of formaldehyde and buildup in OH radicals) occurs slightly earlier for the ethanol-diesel fuel blends. Although, the

12


transition to second stage ignition occurs slightly earlier, the buildup in OH radicals is much more gradual for the E85-diesel fuel blends compared to the gasoline-diesel fuel cases. As previously described, Hashimoto [34] performed detailed kinetics calculations and found that ethanol had a significant inhibitor effect on the cool flame chemistry of heptane mixtures. He attributed this inhibitor effect to consumption of OH radicals by the ethanol molecule. In another work, Hashimoto [42] observed that the ethanol inhibitor effect significantly delayed the second stage heat release of heptane mixtures. The results of Figs. 7 and 8 are consistent with the findings of Hashimoto [[34], [42]]. That is, although significantly higher quantities of diesel fuel (n-heptane in the simulations) were used in the E85-diesel fuel cases compared to the gasoline-diesel fuel cases, the cool flame energy release was significantly reduced. Additionally, the buildup in formaldehyde was delayed for E85diesel fuel blends. Finally, it appears that the presence of ethanol as the diesel fuel transitions to second stage ignition is at least partially responsible for the observed extended combustion duration.

Solid: E85 & Diesel Fuel Dash: Gasoline & Diesel Fuel 0.01

nc7h16 oh ch2o ic8h18 c2h5oh

ic8h18 c2h5oh

Mass Fraction [-]

ic8h18 nc7h16

1E-3

1E-4 ch2o

ch2o

1E-5

oh

-30 -25 -20 -15 -10

-5

oh

0

5

10

15

Crank [° CA ATDC]

Figure 8. Evolutions of n-heptane (nC7H16), iso-octane (iC8H18), ethanol (C2H5OH), formaldehyde (CH2O), and the hydroxyl radical (OH) for E85-diesel fuel blends (solid curves) and gasoline-diesel fuel blends (dashed curves). Note that in the simulations n-heptane is used to represent diesel fuel and isooctane is used to represent gasoline. As previously mentioned, it is thought that the ethanol inhibitor effect is partially responsible for the extended combustion duration observed for E85-diesel fuel blends. However, the dual-fuel charge preparation strategy described in Table 7, results in an in-homogonous fuel species distribution, due to the port-fuel-injection of gasoline or E85 and direct-injection of diesel fuel. That is, stratification in fuel reactivity exits. To understand the influences of the stratified fuel reactivity on the extended heat release for E85-diesel fuel blends, the fuel reactivity distributions of gasoline-diesel and E85-diesel operation were compared. Assigning a research octane number (RON) of 107 to ethanol [37], an overall fuel RON can be defined as

RON =

100 xisooctane + 107 xethanol xisooctane + xnheptane + xethanol

(1)

where x is the mole fraction of each fuel species. Note that, as previously discussed, Hashimoto [34] has shown that ethanol has a significant inhibitor effect on the oxidation of n-heptane; therefore, assigning a reactivity to blends of ethanol and n-heptane may not be as straightforward as presented in Eq. 1. Nevertheless, it is informative to understand the fuel distribution prior to auto-ignition. Figure 9 shows cut-planes shaded by RON and the mass distribution of RON at 20° BTDC, near the time when n-heptane consumption begins due to low-temperature reactions. The mass distribution of RON was calculated by binning the combustion chamber by RON at 20° BTDC. The mass fraction indicates the fraction of the total combustion chamber mass at the specified RON. It can be seen that the gasolinediesel fuel blends have slightly more mass at lower RON than the E85-diesel fuel blends. However, it


13

can also be seen that the range of RON encountered in the combustion chamber for the gasolinediesel fuel blends is lower than that of the E85-diesel fuel blends. That is, the E85-diesel fuel blends have a larger degree of fuel reactivity stratification than the gasoline-diesel fuel blends. Based on the above observations, it appears that the extended combustion duration for the E85-diesel fuel blends is the result of a combination of broader fuel reactivity distributions and the ethanol inhibitor effect described by Hashimoto [34].

Figure 9 Cut-planes on axis coincident with the spray axis colored by RON and the mass distribution of RON at 20° BTDC. The mass fraction shows the fr action of the mass in the combustion chamber at the specified RON. 4.2.3 Mid and High Load Experimental Operation with E85 Exploration of higher engine loads with E85 was desired so as to investigate the large reactivity gradient effect and increased combustion duration at higher loads. As previously described, engine operation with E85-diesel was altered as compared to gasoline-diesel, but the injection timings were not. Similar to operation with gasoline-diesel fueling, increasing load with E85-diesel operation consisted of increasing total fueling and adjusting CA 50 timing through the E85-to-diesel ratio or EGR ratio so as to retain PRR below 10 bar/deg. Throughout the sweep intake temperature was adjusted to 42°C without EGR, and with EGR both intake and EGR cooler temperatures were set to 45°C. This operation strategy demonstrated engine loads between 9.6 and 16.5 bar IMEPg. Where indicated cylinder pressure and AHRR of select crank angles are depicted in Fig. 10. As seen in Fig. 10, increasing load resulted in later combustion phasing, and all loads display extended heat release rates. To compare E85-diesel to previous gasoline-diesel operation, the lowest and highest comparable loads are overlaid in Fig. 11, where the extended combustion duration of E85 at both engine loads is more apparent. Although the 14.2 bar IMEPg operating condition with gasoline-diesel has higher intake pressure than the other loads considered, this condition was at a stoichiometric air fuel ratio at lower loads, thus requiring slightly higher intake pressure to allow additional fueling and load. Regardless of the intake pressure the cases the combustion differences between the two fueling strategies are event. Operation with the gasoline-diesel strategy has a shorter high temperature heat

14


release and more bell-like shape, where as the E85-diesel strategy lends itself to a more triangular shape. It is interesting to note that the trends demonstrated at the mid load condition in Fig. 7, continue at higher engine loads, suggesting similar chemistry. The computational simulation at 9.6 bar IMEPg suggested that the extended combustion duration of E85 occurs from the staged fuel consumption within the reactivity gradient. The staged consumption extends the high temperature combustion duration by first consuming the more reactive diesel fuel before the less reactive gasoline and ethanol fuel components. From the similar measured trends in Fig. 11, and multi peak main heat release trends with various fuel reactivity mixtures and timings [14], it is believed that a similar staged combustion process occurs at higher engine loads. To further examine the differences of E85-diesel and gasoline-diesel operation, the emissions and combustion trends of the respective strategies are plotted in Fig. 12. As seen the CA 50, PRR, and COV between the two strategies are quite similar. Note that the lower COV with increasing load is independent of the fueling strategy. This is believed to be an effect of the higher bulk gas and cylinder wall temperatures at higher engine loads, enhancing end gas combustion and stabilizing initial conditions of the compression process in both residual composition and temperature. This trend was computationally explored with KIVA-CFD predictions that have demonstrated that at lower engine loads significant CO emissions are formed in the end gas and crevice flows. As engine load and thus engine boundary surface temperatures increase, these formation mechanisms are less dominating than the CO oxidation mechanisms. However, as load is continues to increase with both fuel strategies, global equivalence ratios approaching unity are realized. As this occurs, the mid-load optimized in-cylinder charge preparation strategy has areas of higher and lower equivalence ratio, with some areas being fuel rich. It is believed that these slightly rich areas remain sufficiently hot and lean to suppress HC emissions, and prevent significant PM formation, but do result in increased CO emissions. Although HC emissions were measured using the HFID analyzer, the measurements were not corrected for oxygenated species response of the detector. The correction of Kar and Cheng [33] indicates that the HC emissions presented in Fig 12. would increase by approximately 16%, which is not a dramatic difference. Another similar emission trend for gasoline-diesel and E85-diesel operation is NOx emissions, where, as seen in Fig. 12, all tested cases meet EPA 2010 HD limits in-cylinder. Similar NOx emissions suggest that the local temperature histories between the two fueling strategies are similar. These similarities exist even with reduced EGR requirements when port-fuel-injecting E-85. This is believed to be an effect of two differences with E85 operation. First, the combination of enthalpy of vaporization differences of E85 and the additional diesel mass injected resulted in slightly lower average temperatures prior to and throughout combustion. Although intake temperature was adjusted to account for the enthalpy of vaporization differences, the greater temperature gradient from additional injected diesel fuel mass was not accounted for. These temperature effects would in combination extend the combustion duration of both the fuel fraction of diesel and E85. Secondly, the lower reactivity, and thus larger reactivity gradient demonstrated in Fig. 7, would also assist in extension of the E85-diesel combustion duration. The extended E85-diesel combustion duration is not unique to operation without EGR, as is shown in Fig. 7, where similar combustion behavior is observed with 43% EGR and E85 in Fig. 11. From both Figs. 11 and 12, E85-diesel operation is seen to extend the main heat release event by approximately a factor of two. Also in Fig. 12, operation with E85 had significant differences in the EGR requirement as compared to gasoline-diesel operation. Specifically, mid-load E85-diesel operation did not required EGR, while gasoline-diesel required approximately 40%. The decreased EGR requirement of E85diesel operation as compared to gasoline-diesel offers additional benefits beyond different air handling requirements. For example, all E85-diesel cases displayed increased thermal efficiency. Figure 12 demonstrates these trends where the minimum ISFC of E85-diesel was found to be 142 g/kW-hr on a gross gasoline equivalent basses (189 g/kW-hr not correct for gasoline equivalence), and a corresponding gross indicated thermal efficiency of 59.0%. The additional efficiency gains with E85-diesel are primarily from raised specific heat benefits with lower or no EGR operation at all tested loads, as compared to gasoline-diesel. Similar to high load gasoline-diesel operation, further optimization of high-load E85-diesel operation might provide additional efficiency at the higher loads.


160

Pressure (bar)

120

1.6

NTC Behavior

AHRR (kJ/° CA)

140

9.6 bar IMEPg | 11.6 bar IMEPg 14.9 bar IMEPg | 47% EGR 16.5 bar IMEPg

100

0.05 0.04 0.03 0.02 0.01 0.00 -20 -15 -10 -5 0 Crank Angle (° CA ATDC)

1.4 1.2 1.0

80

0.8

60

0.6

40

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20

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AHRR (kJ/° CA)

0% EGR 43% EGR

15

0.0

-25 -20 -15 -10

-5

0

5

10

15

20

25

30

Crank Angle (° CA ATDC) Figure 10 Measured E85-diesel for loads from 9.6 to 16.5 bar IMEPg. Combustion phasing was varied to retain PRR below 10 bar/deg Gasoline Diesel E-85/Diesel 160

9.6 bar IMEPg 37% EGR| 9.6 bar IMEPg 0% EGR |

14.2 bar IMEPg 57% EGR 14.9 bar IMEPg 43% EGR 1.6

Pressure (bar)

120 100

1.4

0.04 0.03 0.02

1.2

0.01 0.00 -20

-15

-10

-5

Crank Angle (° CA ATDC)

0

1.0

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40

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-25 -20 -15 -10

-5

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10

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AHRR (kJ/° CA)

140

AHRR (kJ/° CA)

NTC Behavior 0.05

30

Crank Angle (° CA ATDC) Figure 11 Measured cylinder pressure and AHRR at 9.6 and ~14 bar IMEPg loads with E85-diesel and gasoline-diesel operation. Both loads display low PRR and long HTHR duration, with E85-diesel displaying extended HTHR as compared to gasoline-diesel.

20 18 16 14 12 10 8 6

0.015

6

4.0 3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0

0.3 0.2

2010 HDlimit 0.1 0.0 4

6

8

10

12

14

16

COV (%)

2

12 10 8 6 4 2 0 -2

ISFCg gasoline eq.(g/kW-hr)

0.000

10

180

η g indicated (-)

NOx (g/kW-hr)

PRR (bar/° CA)

0.005 14

CA50 (° CA ATDC)

0.010

18

60 50 40 30 20 10 0

8 7 6 5 4 3 2 1 0

16

170

12

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8

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4

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0.60 0.58 0.56 0.54 0.52 0.50 0.48 0.46

IMEPg (bar)

E-85/diesel gasoline/diesel

0

1.0

CO (g/kw-hr)

PM (g/kW-hr)

2010 HD limit

HTHR 5-90 (° CA)


0.020

HC (g/kw-hr)


EGR (%)


0.8

Φ (−)

16

0.6 0.4 0.2 0.0

4

6

8

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16

18

4

6

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IMEPg (bar)

Figure 12 Measured combustion and emissions load trends with E85-diesel and gasoline-diesel. Note the high thermal efficiency and low emissions of both cases. Also, note the approximately double HTHR duration with E85 as compared to gasoline

5. Summary and Conclusions The presented experimental and computational results demonstrate that fuel reactivity controlled combustion is a viable LTC strategy throughout the typical load range of heavy duty engines. The incylinder blended dual fuel strategy has been demonstrated to offer high thermal efficiency and low emissions, with both conventional and renewable fuels. Gasoline-diesel operation was demonstrated at engine loads up to 14.5 bar IMEPg, and E85-diesel operation was successful at engine loads as high as 16.5 bar IMEPg, with neither strategy being load limited by engine pressure or combustion constrains. For all tested conditions E85-diesel operation was shown to exhibit significant combustion differences as compared to gasoline-diesel operation. The extended combustion duration was found to be a function of temperature, local and global reactivity differences, and chemistry differences with ethanol-diesel fuel blends. These differences with E85-diesel as compared to gasoline-diesel fueling enabled lower rates of EGR at all engine loads. The lower EGR rates were found to be beneficial to increasing thermal efficiency, with the maximum efficiency of E85 diesel measured to be 59% and gasoline-diesel to be 56%, both on a gross indicated basis. All tested conditions with E85-diesel operation were found to meet EPA 2010 heavy duty emissions mandates for NOx and soot in-cylinder, with gasoline-diesel operation also meeting NOx and soot at mid loads. At higher loads, gasoline-diesel operation attained a stoichiometric equivalence ratio, resulting in soot emissions slightly over EPA 2010 mandates. However, fuel consumption and emissions remained low for stoichiometric CI operation. This study has demonstrated that with an injection strategy optimized for mid-load, good results are possible at low and high load operation, confirming that the charge preparation strategy is robust [43]. It has also been demonstrated that fuel effects within the reactivity gradient are responsible for controlling the combustion event. Further experimental and computational explorations at higher and lower loads, with different fuels, and EGR effects will be investigated with fuel reactivity-controlled dual fuel combustion in future tests.

Acknowledgements The authors thank the United States Department of Energy through grant DE-EE0000202, and the University of Wisconsin Engine Research Center Diesel Emissions Reduction Consortium (DERC) for funding, as well as United Wisconsin Grain Producers for supplying the ethanol used in the research.


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References [1]

[2]

[3]

[4] [5]

[6] [7]

[8] [9] [10]

[11]

[12]

[13]

[14]

[15]

[16]

[17] [18] [19]

[20]

[21] [22]

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. Dec J.E. (2003) Diesel-fueled HCCI engines. In: Zhao F, Asmus T.A., Assanis D.N., Dec J.E., Eng J.E., Najt P.A. (eds) Homogeneous Charge Compression Ignition (HCCI) Engines. SAE International, Warrendale, PA, 2003. Kimura S, Aoki O, Kitahara Y, Aiyoshizawa E (2001) Ultra-Clean Combustion Technology Combining a Low-Temperature and Premixed Combustion Concept for Meeting Future Emission Standards. SAE paper 2001-01-0200, 2001. Kimura S, Aoki O, Ogawa H, Muranaka S, Enomoto Y (1999) New Combustion Concept for Ultra-Clean and High-Efficiency Small DI Diesel Engines. SAE paper 1999-01-3681, 1999. 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. Hanson, R.M., Splitter, D.A., Reitz, R.D., “Operating a Heavy-Duty Direct-Injection Compression-Ignition Engine with Gasoline for Low Emissions”, SAE paper 2009-01-1442, 2009. 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. Kamimoto T., and Bae M., "High Combustion Temperature for the Reduction of Particulate in Diesel Engines", SAE Technical Paper 880423, 1988. Sun Y., Reitz R.D., “Modeling Diesel Engine NOx and Soot Reduction with Optimized TwoStage Combustion”, SAE Technical Paper 2006-01-0027, 2007. Akihama K., Takatori Y., Inagaki K., Sasaki S. and Dean A.M., "Mechanism of the Smokeless rich Diesel Combustion by Reducing Temperature", SAE Technical Paper 2001-01-0655, 2001. Bessonette P.W., Schleyer C.H., Duffy K.P., Hardy W.L., and Liechty M.P., "Effects of Fuel Property Changes on Heavy-Duty HCCI Combustion", SAE Technical Paper 2007-01-0191, 2007. 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. Kokjohn S.L., Hanson R.M., Splitter D.A., and Reitz R.D., "Experiments and Modeling of DualFuel HCCI and PCCI Combustion Using In-Cylinder Fuel Blending", SAE Technical Paper 2009-01-2647", 2009. 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. 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. 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. Nevin R.M., “PCCI Investigation Using Variable Intake Valve Closing in a Heavy Duty Diesel Engine”, M. S. Thesis, University of Wisconsin – Madison, 2006. Amsden A.A., "Kiva-3v, release 2, improvements to kiva-3v", LA-UR-99-915, 1999. Abani N., Kokjohn S.L., Park S.W., Bergin M., Munnannur A., Ning S.Y.W., and Reitz R.D., "An Improved Spray Model for Reducing Numerical Parameters Dependencies in Diesel Engine CFD simulations", SAE Technical Paper 2008-01-0970, 2008. Abani N., Munnannur A., and Reitz R.D., "Reduction of Numerical Parameter Dependencies in Diesel Spray Models", Journal of Engineering for Gas Turbines and Power, vol.130, no.3 2008. Han Z. and Reitz R.D., "Turbulence Modeling of Internal Combustion Engines Using RNG k-e Models", Combustion Science and Technology, vol.106 1995. Ra, Y. and Reitz, R.D. "A Combustion Model for IC Engine Combustion Simulations with MultiComponent Fuels", submitted to Combustion and Flame, 2009.

18


[23]

Ra Y., Yun J.E., and Reitz R.D., "Numerical Parametric Study of Diesel Engine Operation with Gasoline", Combustion Science and Technology, vol.181, pp. 1-29 2008. Kong S.C., Sun Y. and Reitz R.D., "Modeling Diesel Spray Flame Lift-off, Sooting Tendency and NOx Emissions Using Detailed Chemistry with Phenomenological Soot Model", ASME Journal of Gas Turbines and Power, vol.129, pp. 245-251 2007. Hiroyasu H., and Kadota T., "Models for Combustion and Formation of Nitric Oxide and Soot in DI Diesel Engines", SAE Technical Paper 760129, 1976. Sun Y., "Diesel Combustion Optimization and Emissions Reduction Using Adaptive Injection Strategies (AIS) with Improved Numerical Models", PhD Thesis, University of Wisconsin-Madison, 2007. Frenklach M., Bowman T., Smith G., and Gardiner G., "Http://www.Me.Berkeley.Edu/grimech/". Abraham J., "What is Adequate Resolution in the Numerical Computation of Transient Jets", SAE paper 970051, 1997. Beale J.C., and Reitz R.D., "Modeling Spray Atomization with the Kelvin-Helmholtz/RayleighTaylor Hybrid Model", Atomization and Sprays, vol.9, pp. 623-650 1999. Munnannur A., "Droplet Collision Modeling in Multi-Dimensional Engine Spray Computations", PhD. Thesis, University of Wisconsin-Madison, 2007. O'Rourke P.J., and Amsden A.A., "A Particle Numerical Model for Wall Dynamics in PortInjected Engines", SAE Technical Paper 961961, 1996. O'Rourke P.J. and Amsden A.A., "A Spray/wall Interaction Submodel for the KIVA-3 Wall Film Model", SAE Technical Paper 2000-01-0271, 2000. 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. Hashimoto K., “Inhibition Effect of Ethanol on Homogeneous Charge Compression Ignition of Heptane”, SAE Technical Paper 2008-01-2504, 2008. Marinov, N.M., “A Detailed Chemical Kinetic Model for High Temperature Ethanol Oxidation”, International Journal of Chemical Kinetics 31, 183-220, 1999. 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). Heywood J.B., "Internal Combustion Engine Fundamentals" McGraw-Hill, New York ,1988 Egolfopoulus F.E., DU D.X., Law C.K., “A Study on Ethanol Oxidations Kinetics in Laminar th Premixed Flames, Flow Reactors, and Shock Tubes” 24 symposium (international) on Combustion/ the Combustion Institute,1992, pp 833-841. Röhl O., Jerzembeck S., Beeckmann J., and Peters N., "Numerical Investigation of Laminar Burning Velocities of High Octane Fuel Blends Containing Ethanol" SAE Technical Paper 2009-01-0935, 2009. Virorrio M., Tuenstal P., Johansson B., " Effects of Ethanol and Different Type of Gasoline Fuels on Partially Premixed Combustion from Low to High Load ", SAE Technical paper 201001-0871, 2010 Warnatz J., Maas U., Dibble R.W., "Combustion Physical and Chemical Fundamentals, Modth eling and Simulation, Experiments, Pollution Formation", 4 edition, Springer Berlin Heidelber, New York, 2006. Hashimoto K., "Effect of Ethanol on the HCCI Combustion", SAE Technical Paper 2007-012038, 2007. 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

[24]

[25] [26]

[27] [28] [29] [30] [31] [32] [33] [34] [35] [36] [37] [38]

[39]

[40]

[41]

[42] [43]