Title: The application of diesel-fueled HCCI combustion in an automotive-size single-cylinder diesel engine

Choongsik Bae*, Sanghoon Kook and Seik Park

Engine Laboratory, Department of Mechanical Engineering, Korea Advanced Institute of Science and Technology (KAIST), 373-1 GuSeong-Dong, YuSeong-Gu, Daejeon 305-701, Republic of Korea

E-mail: csbae@kaist.ac.kr (Choongsik Bae) Telephone: +(82) 42 869 3044 Fax: +(82) 42 869 5044

Submitted to the Second International Symposium on “Clean and High Efficiency Combustion in Engines”, July 10-13 2006, Tianjin, China

* Corresponding author

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The Application of Diesel-Fueled HCCI Combustion in a Single-Cylinder Diesel Engine

Choongsik Bae, Sanghoon Kook1 and Seik Park Engine Lab., Dept. of Mechanical Eng., KAIST

ABSTRACT

The operation of a diesel-fueled homogeneous charge compression ignition (HCCI) engine was studied in a single-cylinder direct-injection optical diesel engine with regard to several key parameters: spray penetrations, the dilution of the intake air, and two-stage fuel injection. The influence of these parameters on HCCI combustion was clarified through spray measurements and combustion analysis. The spray penetrations were optimized by a small included angle to avoid wall-impingement at low pressure and temperature in the cylinder. Analysis of the ignition delay shows that there is minimum premixing time to guarantee low combustion temperature regardless of the engine speed. Dilution of the intake air by adopting exhaust gas recirculation (EGR) retards the ignition timing thereby controlling the combustion phasing. Furthermore, applying the two-stage injection such that a small amount of the ignition promoting fuel which is injected near TDC to assist the combustion of a premixed charge results in the most applicable HCCI combustion in a diesel engine. KEYWORDS: Homogeneous charge compression ignition (HCCI), Direct-injection, Spray wall-impingement, Small included angle, Injection timing, Exhaust gas recirculation, Two-stage injection 1. INTRODUCTION A diesel-fueled homogeneous charge compression ignition (HCCI) combustion is an advanced technique for reducing hazardous nitrogen oxides (NOx) and particulate matter (PM) emissions at low load operation, while providing high fuel conversion efficiency in a diesel engine. NOx emission can be reduced by achieving lean homogeneous mixture through the entire cylinder volume, resulting in low combustion temperature. A homogeneous charge having no fuel-rich zones, unlike conventional diesel combustion, reduces PM emission as well. There are two main approaches to supply a homogeneous charge before ignition: A port-injection where diesel fuel is supplied to the intake port [1, 2], and an in-cylinder injection where fuel is injected directly into the combustion chamber. Although port-injection theoretically guarantees sufficient time for premixing and enables more successful HCCI combustion, wall-wetting on account of the low temperature condition in the intake port is a major problem in actual applications. Moreover, as the HCCI combustion has limited operation range, the mode transition between HCCI and conventional diesel combustion is required that is not preferable with port-injection system [3]. For better engine application, in-cylinder direct-injection has been widely tested with the most advanced common-rail fuel injection system [4 – 12]. The accomplishment of HCCI engine combustion with the in-cylinder direct-injection method is particularly influenced by following parameters:

1. Spray penetrations: The fuel spray penetrations should be long enough to achieve spatially uniform fuel distribution, but simultaneously, short to impede the spray-

1 Currently with the Korea Institute of S&T Evaluation and Planning (KISTEP)

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impingement to the cylinder liner or the piston. This parameter is controlled by the injector geometry. Multi-hole injectors have been applied to promote the atomization and spatial distribution of the mixture [4, 5]. A swirl injector with a pintle type nozzle and two impinging side-injectors have been tested to minimise the spray wall-impingement [6]. Also an impinged-spray nozzle injector [7], or a hole-injector with a small included angle [8, 9] were investigated to resolve the same problem.

2. The dilution of the intake air: Generally, the ignition timing of HCCI combustion is earlier than that of conventional diesel combustion because of the hot combustion driven by the cool combustion [1, 3, 13]. This early ignition problem becomes more severe if the intake air is preheated to assist vaporization of the diesel [1, 2, 6]. Previous researchers applied the charge dilution as employing EGR to retard the combustion phase [2, 6, 7, 8, 10]. Diluting the intake charge extends the ignition delay and lowers the combustion temperature due to the reactant being replaced by the inert gases with increased heat capacity.

3. Two-stage injection: Some studies employed multiple injection strategies using only one injector [4, 9, 10] or several injectors [5, 6]. The earlier injection was performed to form a premixed charge without the spray-impingement problem, and was followed by the injection near the top dead center (TDC). Other studies, meanwhile, have focused on injections near TDC, while pre-mixing is promoted by high swirl ratio and long ignition delay employed by a high rate of EGR [11, 12].

The objective of this study is to suggest engine operating strategies regarding the parameters noted above. Macroscopic spray visualisation were conducted for a sac nozzle injectors with various included angles (70 ~ 150°) and hole numbers (5 ~ 14 holes) in a spray chamber. The fuel injection timing was widely controlled from the compression stroke to the intake stroke with a common-rail fuel injection system incorporated in a single-cylinder diesel engine at both 800 rpm and 1200 rpm. Also the flame images were obtained over a wide range of injection timing. EGR was employed up to 40%, so as to control combustion phase. In addition, a small amount of fuel (1.5 mm3) was injected near TDC to promote ignition of in-cylinder charge with preheated intake air of 160°C. Its effect on power-output and combustion stability was investigated over various fuel quantities. 2. EXPERIMENTS

2. 1 Spray measurements in a chamber

The spray penetrations of sac nozzle injectors were measured in a spray chamber with a volume of 13,804 cm3 under atmospheric pressure and temperature. Figure 1 shows the experimental setup. The chamber allows optical access through circular windows 89 mm in diameter. Spray images frozen by a spark light source (EG&G Optoelectronics; MVS-2601-CE96), which had light duration of shorter than 100 ns, were acquired with a CCD camera (PCO CCD Imaging; SensiCam). The macroscopic spray structures were manifested by this imaging technique [14]. The air density or pressure was controlled to meet the conditions at early injection timing, generally an open-valve condition, which is close to atmospheric pressure. At first, the 5-hole injector (Bosch) with included angle of 150˚ was tested, and then smaller included angles of 100˚ and 70˚ were applied to overcome spray wall-impingement on the cylinder liner. Injectors of 8- and 14-holes were tested to verify the enhancement of the atomization as well as shorten the spray penetration. The main injector geometries are listed in Table 1. In case of spray-targeting visualization, the radial cut piston was inserted in a chamber.

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Fig. 1 Schematic diagram of spray chamber setup

Table 1 Injector geometries (sac nozzle injectors; Bosch)

Case Hole number Hole diameter Included angle I 5 0.168 mm 150˚ II 5 0.168 mm 100˚ III 8 0.133 mm 100˚ IV 14 0.100 mm 150˚ V 14 0.100 mm 100˚ VI 14 0.100 mm 70˚

Table 2 Specifications of the research engine

Engine type single-cylinder, direct injection, four-valves, optical diesel engine Bore 83 mm

Stroke 92 mm Displacement 498 mm3

Compression ratio 14.8, 18.7 Fuel Injection

Equipment Bosch common-rail

2. 2 Research engine The single-cylinder optical HCCI engine (Engine Tech; RSi-090) used for this study is based on a typical direct-injection diesel engine [9]. The bore and stroke were selected to be comparable to the specifications of a typical passenger car engine. Its specifications are summarized in Table 2. A schematic diagram of engine setup is shown in Fig. 2 (see the next page). The temperature of the intake air was controlled by an electrical heater and k-type thermocouple with a temperature control unit. Fuel injection parameters including injection pressure, quantity, and timings were controlled by a programmable injector driver (TEMS; TDA3000H) and a pressure regulator. A rotary encoder (Koyo; 3600 pulses/revolution) attached to a camshaft was used to control the operating timings of the multiple fuel injections and the camera. In-cylinder pressure was recorded with a piezoelectric pressure transducer (KISTLER; 6052A) at every 0.16 crank angle degrees (CAD). A smokemeter (AVL) was used to measure the engine-out soot emission. An emission measurement system (HORIBA; MEXA1500D) sampling the exhaust gas was employed to measure the concentrations of NOx, HC, and CO. An EGR line was installed from the exhaust pipe to the intake pipe with a surge tank to minimise the pressure fluctuations. For higher level of charge dilution, supplemental CO2 was supplied through the intake pipe. The cylinder pressures and exhaust gas concentrations of 138 engine cycles were recorded and averaged to calculate indicated mean effective pressure (IMEP) and apparent heat release rate.

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Fig. 2 Schematic diagram of the engine setup and visual field

Figure 2 also illustrates the optical access to the combustion chamber and observation

field. An elongated piston was installed in order to enable the mounting of a 45˚ mirror beneath the piston quartz window, which allowed a 50 mm diameter visual field in 83 mm bore. A high speed digital video camera (Vision Research; Phantom v7.0) with an image intensifier (PROXITRONIC) was used to acquire the combustion images inside the optical engine. This setup was corresponding to the high speed imaging up to 10,000 frames per second with resolution of 512 x 384 pixels, thus the images could be taken each 0.48 CAD. The exposure time of the camera was optimized to 35 s to obtain flame imagesμ without blurring. The images were highly amplified through an intensifier to show the non-luminous flame. The maximum gain of the intensifier was 108 W/W and the maximum light power per area was 70 μW/cm2. An intensity gain of 50% was used for detecting weak HCCI combustion except in the highly luminous case where the gain value was decreased to 30%. 2. 3 Engine operating conditions The engine operating conditions are listed in Table 3. The engine was operated at 800 and 1200 rpm under both motored and fired conditions. In order to maintain consistency of performance data and flame images, peak motored pressure was kept at 3.2 MPa for every engine operating cycles. This was achieved by controlling the motoring period before fuel injection. During the motoring, the peak pressure kept rising and then reached stable condition in 120 s. The coolant temperature was controlled at 353 K and diesel fuel temperature at 313 K. The common-rail pressure that was close to the injection pressure was selected as 120 MPa.

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This high pressure injection could aggravate the wall-impingement problem. However, our

earlier study showed that the total heat release was increased with increasing injection pressure due to the improved atomization and the better mixing with air [9]. Therefore, the injection pressure was maintained at a relatively high value in order to achieve as high power-output as possible, which was comparable to that in the conventional diesel engine. The intake air temperature was controlled and held constant at 160°C. This temperature was high enough to assist the vaporization of diesel fuel, such that the power-output was high even with the early injection [9]. In this study, the early ignition problem was resolved by introducing EGR. The EGR rate is defined as the ratio between CO2 concentration of the intake air and CO2 concentration of the exhaust concentration. EGR rate was varied from 0 to 40%. The fuel quantity was varied from 11.5 to 16 mm3. The injection timings were widely changed from 50 CAD before top dead center (BTDC) to 200 CAD BTDC.

Table 3 Engine operating conditions.

Engine speed 800, 1200 rpm Coolant temperature 353 K

Diesel fuel temperature 313 K Common-rail pressure 120 MPa Intake air temperature 433 K

EGR rate 0 ~ 40 % Main injection timing 200 ~ 50 crank angle degree (CAD) BTDC

Second injection timing of the two-stage injection

20, 10 CAD BTDC and TDC

Total quantity of fuel injection 11.5 ~ 16 mm3

Quantity of second injection 1.5 mm3

Fuel Conventional diesel (cetan no.=50)

Fig. 3 The macroscopic spray visualisation in the chamber; 14-hole injector, included

angle=100° (Case V)

3. RESULTS AND DISCUSSION 3. 1 Spray penetrations The spray penetrations were measured from macroscopic liquid spray images in the spray chamber. Figure 3 shows an example. Because the line of sight of camera was longitudinal to the injector axis, it depicts radial penetrations for the 14-hole injector. The position of the cylinder wall is superimposed. The 10 images were averaged to estimate the penetration. This procedure was repeated for each hole.

A small included angle, extending the axial spray penetration and shortening the radial spray penetration, was examined to give guidance to the engine designers by clarifying the

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spray wall-impingement. Figure 4(a)~(c) shows the axial and radial spray penetrations relative to the position of the piston and cylinder liner at various included angles both for 5-hole and 14-hole injectors. For the 5-hole injector (Fig. 4(b)), the penetration of conventional diesel injection was shown under a high charge air density of 33.8 kg/m3. The spray wall-impingement problem occurred under low charge air density of 1.189 kg/m3. The axial development of spray ( z ) did not impinge on the top of the piston if the injection timing was controlled near BDC, but the radial development of spray ( r ) predicted severe wall-impingement. The spray impinged the cylinder liner if the included angle was 150°. The increased wetting could cause poor mixing and air utilization resulting in low combustion efficiency and high PM emission. The spray with the included angle of 100° impinged less than the included angle of 150° did. The confliction could be more resolved with the included angle of 70° as shown in Fig. 8(c) for the 14-hole injector.

S

S

(a) Definition of included angle

100

80

60

40

20

00 10 20 30 40 50 60

HCCI

5 holes,IncludedAngle;

150O (Case I) 100O (Case II)

Charge Air Density=1.189 kg/m3 (HCCI )=33.8 kg/m3 (DI-diesel)

Radial Spray Penetration(Sr) [mm]

Piston Top at BDC

Cylinder Liner

Axi

al S

pray

Pen

etra

tion(

Sz) [

mm

] DI-diesel

(b) Penetrations with 5-hole injector

100

80

60

40

20

00 10 20 30 40 50 60

Radial Spray Penetration(Sr) [mm]

Piston Top at BDC

14 holes,Included Angle;

150O (Case IV) 100O (Case V) 70O (Case VI)

Charge Air Density =1.189 kg/m3 (HCCI )

Cylinder Liner

Axi

al S

pray

Pen

etra

tion(

Sz)

[mm

]

(c) Penetrations with 14-hole injector

Fig. 4 Effect of included angle on spray wall-impingement; 5- and 14-hole injectors

It has been reported that application of a multi-hole injector (e.g. 30 holes, 0.100 mm hole diameter) for HCCI combustion promotes atomization of the spray [4]. It was also reported that supplying higher injection pressure through a smaller hole facilitates better fuel atomization and evaporation [18], and the smaller hole diameter tended to increase air entrainment into the spray [19]. On the other hand, some researchers found that small hole diameter reduced spray penetration, resulting in poor spatial-distribution of fuel [20] and created a fuel-rich region [21]. As the mode transition with only one fueling system is one of the advantages of the early direct-injection method, poor air utilization for the conventional diesel combustion mode, corresponding to warm-up, cold-start, or high load condition, would be a serious source for high PM emission. From this point of view, an overly small included angle would also be a problem for near TDC injection. As we focused on HCCI combustion rather than the mode transition in this study, however, the 14-hole injector with an included angle of 70° (Case VI) was used to have more investigation through the following experiments. 3. 3 Dilution of the intake air

One of the significant problems in application of HCCI engine combustion is the early

ignition resulting in low fuel conversion efficiency on account of the ineffective combustion

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phasing. Previous researchers focused on the effect of dilution of the intake air as employing EGR to retard the ignition [6 - 8].

Fig. 11 Effect of exhaust gas recirculation on in-cylinder pressure trace and heat release rate;

single-injection, injector Case VI, intake air temperature=160°C, injection pressure=120 MPa

12 13 14 15 1605

10152025303540

Combustion efficiency

Fuel quantity (mm3)

EGR

rate

(%)

88.0088.6389.2589.8890.5091.1391.7592.3893.00

12 13 14 15 16

05

10152025303540

Fuel quantity (mm3)

Work conversion efficiency

EGR

rate

(%)

38.5039.1939.8840.5641.2541.9442.6343.3144.00

(a) Combustion efficiency [%] (b) Work conversion efficiency [%]

12 13 14 15 1605

10152025303540 Fuel conversion efficiency

Fuel quantity (mm3)

EGR

rate

(%)

20.0020.4420.8821.3121.7522.1922.6323.0623.50

(c) Fuel conversion efficiency [%]

Fig. 12 Effect of EGR rate and injected fuel quantity on combustion efficiency, work conversion efficiency, and fuel conversion efficiency; single injection, injection pressure=1200bar, intake air temperature=160˚C, compression ratio=14.8, injector Case VI, engine speed=1200 rpm

Figure 11 shows that the ignition is retarded as the EGR rate increases from 0% to

40%. The peak of the heat release rate decreased, but the combustion period extended resulting in high fuel conversion efficiency. It was explicit that the fresh air in the intake pipe was diluted by the exhaust gas and, due to the reactant being replaced by the inert gases with

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increased heat capacity, the chemical reaction of the premixed charge was slowed down. Figure 12(a) shows the combustion efficiency, work conversion efficiency, and fuel conversion efficiency over the injected fuel quantity and EGR plane. As an aid to understanding the influence of various factors on fuel conversion efficiency fcη , we have previously proposed [24] that fcη be decomposed as a product of three separate efficiencies

4847644844764484476 chlwc ηηη

⎞⎛

LHVf

chem

chem

hlchem

hlchemfc Qm

QQ

QQQQ

Wη ⎟⎟⎠

⎜⎜⎝⎟⎟⎠

⎞⎜⎜⎝

⎛ −⎟⎟⎠

⎞⎜⎜⎝

⎛−

=

The equivalence ratio is getting richer in the direction of the upper-right corner and

leaner in the direction of the lower-left corner. The combustion efficiency cη was estimated from the HC and CO emissions (more details can be found in Ref. 24). The combustion efficiency decreased at high EGR rate at fixed fuel quantity because the combustion temperature was decreased on account of the diluents in the mixture. As more fuel was injected at fixed EGR rate, the combustion efficiency was increased because the flame temperature was increased so that HC/CO emissions were decreased as the equivalence ratio was getting closer to the stoichiometric (recall the overall equivalence ratio is under 1 for the diesel engine). The work conversion efficiency wcη – the work produced over chemical energy released minus heat loss – was calculated from the in-cylinder pressure data. As the combustion phasing was retarded with the increasing EGR rate at fixed fuel quantity, work conversion efficiency was increased. Another finding from the Fig. 12(b) was that the work conversion efficiency increased when the equivalence ratio was higher within the operating range of this study. It was anticipated that the lean mixture induced the faster combustion phasing as the less endothermic energy was required with small amount of fuel. The fuel conversion efficiency fcη is proportional to the combustion efficiency and the work conversion efficiency. As a result, figure 12(c) shows that the fuel conversion efficiency is highest at mid-level of EGR rate at fixed fuel quantity and it is increasing with the fuel quantity. It was the result of trade-off between the decreasing combustion efficiency and increasing work conversion efficiency with increasing EGR rate. Note that the drawback in applying EGR to HCCI combustion is transient response and temperature instability of recycled gas [7]. The coefficient of variation (COV) of engine work was 3% or less in the presented results.

3. 4 Two-stage injection

Previously, the two-stage injection strategy was introduced as a concept of ignition-enhancing to solve the problem in low power-output [9]. It was achieved by injecting small amount of fuel (1.5 mm3 ) termed as second injection near TDC, while most of the fuel (10 mm3) termed as main injection was injected earlier than 100 CAD BTDC to make a lean homogeneous charge. In this study, this strategy was more precisely tested in selecting the second injection timing. Figure 13 shows the tested second injection timing; ‘a’ is at the start of high-temperature combustion, ‘b’ is during the very active premixed combustion, ‘c’ is at the middle of declination of combustion, and ‘d’ is after the combustion. At ‘a’, the second injection was expected to play a role of ignition promotion, and from ‘b’ to ‘d’, the burn-out of remained HC/CO was predicted at each combustion stages.

Figure 14 shows the IMEP, HC/CO, NOx and smoke emissions for each of the second injection timing along with the single injection. The injection timing of the single injection was 70 CAD BTDC (optimal fuel distribution due to the spray-targeting (see Fig. 9 and 10). The total fuel quantity was maintained as 16 mm3 regardless of the injection strategies such that 16 mm3 was injected at once for the single injection, and for the two-stage injection, 14.5 mm3 was injection first and then the 1.5 mm3 of second injection was followed. Overall, figure 14 shows the advantages of the two-stage injection strategy. Compared to the single injection, IMEP was increased over 20% and significant decrease in HC/CO emissions was captured. However, NOx and smoke emissions were increased depending on the second injection timing. The increase of the flame temperature as supplying the other heat source resulted in significant decrease of

9

-5

0

5

10

15

20

25

30

-40 -20 0 20 40

d

c

b

a

Hea

t rel

ease

rate

(J/d

eg)

Crank Angle (CAD ATDC)

Fig. 12 The selection regime of the second injection timing on the heat release rate curve

Fig. 13 The effect of second injection timing on power and emissions; main injection=70 CAD

BTDC, total injection quantity=16 ㎣, second injection quantity=1.5 ㎣, injection pressure=120 MPa, intake air temperature=160˚C, compression ratio=14.8, injection angle=70˚, engine speed=1200 rpm

the HC/CO but simultaneously the increase of the soot and NOx formations were followed. Concerning the selection of the second injection timing, the tremendous increase of

NOx emission was notable at ‘a’ which explained the thermal assist of the high-temperature ignition would result in increase of flame temperature. The enhancement of the combustion was obvious at ‘b’ such that the highest IMEP and lowest HC/CO emissions. However, the second highest NOx emission was a problem. The second injection was effective at ‘c’ when it acted as the burn-out source of the HC/CO emissions while the flame temperature was kept low enough to hinder the soot and NOx formations.

Analysis of combustion efficiency, work conversion efficiency, and fuel conversion efficiency was repeated for the two-stage injection. Figure 15 shows the same trends with single injection (Fig. 12) in combustion efficiency and work conversion efficiency while more EGR gas was needed to achieve maximum fuel conversion efficiency.

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12 13 14 15 1605

10152025303540

Combustion efficiency

Fuel quantity (mm3)

EGR

rate

(%)

85.0086.0087.0088.0089.0090.0091.0092.0093.00

12 13 14 15 1605

10152025303540

Fuel quantity (mm3)

EGR

rate

(%)

Work conversion efficiency

44.5045.0045.5046.0046.5047.0047.5048.0048.50

(a) Combustion efficiency (b) Work conversion efficiency

12 13 14 15 1605

10152025303540

Fuel quantity (mm3)

EGR

rate

(%)

Fuel conversion efficiency26.0027.3828.7530.1331.5032.8834.2535.6337.00

(c) Fuel conversion efficiency

Fig. 14 Effect of injected fuel quantity and EGR rate on combustion efficiency, work conversion efficiency, and fuel conversion efficiency; two-stage injection, second injection timing = ‘c’, injection pressure=120 MPa, intake air temperature=160˚C, compression ratio=14.8, injector Case VI

4. CONCLUSIONS Several key parameters – spray penetrations, the dilution of the intake air, and two-stage fuel injection influencing HCCI combustion – were tested in a single-cylinder automotive-size direct-injection diesel engine. From the spray/flame imaging and thermodynamic analysis of the in-cylinder pressure data, the major findings in terms of the engine application of the HCCI combustion can be summarized as follows: 1. An injector with a small included angle of 70° can reduce the spray wall-impingement by

extending the axial spray penetration and shortening the radial spray penetration. 2. Due to the trade-off between combustion efficiency and work conversion efficiency with

increasing EGR rate from 0 to 40%, maximum fuel conversion efficiency was observed at mid-level of EGR rate.

3. Two-stage injection with a small amount of the ignition promoting fuel (1.5 mm3), which was injected near TDC to assist the combustion of a premixed charge resulted in IMEP increase and HC/CO decrease. The optimal second injection timing was found at the middle of declination of premixed combustion such that high IMEP, low HC/CO, and

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reasonable smoke and NOx emissions. ACKNOWLEDGEMENT

Support for this research was provided by the CERC (Combustion Engineering

Research Center) and Future vehicle technology development corps of Korea. The research was performed at the engine laboratory of the department of mechanical engineering, KAIST. KAIST is a research-oriented educational institute operated by Korea Ministry of Science and Technology. Thanks are also due to Paul Miles of Sandia National Laboratories for his original idea in analyzing the combustion efficiency, work conversion efficiency, and fuel conversion efficiency.

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13. Yanagihara, H. A Study on Combustion Structure of Premixed Compression Ignition Diesel Engines, Proc. of THIESEL 2002 Conference on Thermo- and Fluid-Dynamic Processes in Diesel Engines, Homogeneous Charge C. I., 2002, pp.111-129 14. Kang, J., Bae., C., and Lee., K. O. Initial development of Non-evaporating Diesel Sprays in Common-rail Injection Systems, International Journal of Engine Research, 2003, Vol. 4, No. 4, pp283-298 15. Bosch, W. The fuel rate indicator: a new measuring instrument for display of the characteristics of individual injection, SAE Technical paper 660749, 1966. 16. Arai, M., Tabata, M., Hiroyasu, H. and Shimizu., M. Disintegrating process and spray characterization of fuel jet injected by a diesel nozzle, SAE Technical Paper 840273, 1984 17. Naber, J. D. and Siebers, D. L. Effects of gas density and vaporization on penetration and dispersion of diesel spray, SAE Technical Paper 960034, 1996 18. Heywood, J. B. Internal Combustion Engine Fundamentals, McGraw-Hill, Inc., 1988 19. Siebers, D. L. and Higgins, B. S. Flame Lift-Off on Direct-Injection Diesel Sprays Under Quiescent Conditions, SAE Technical Paper 2001-01-0530, 2001 20. Hiroyasu, H. and Arai, M. Structures of Fuel Sprays in Diesel Engines, SAE Technical Paper 9000475, 1990 21. Rente, T., Gjirja, S., Denbratt, I. Experimental Study On The Influence Of Nozzle Orifice Number On Combustion And Emissions Formation Of A Heavy Duty D.I. Diesel Engine, SAE_NA Technical Paper 2003-01-67, 2003 22. Christensen, M., Johansson, B., and Hultqvist, A. The Effect of Topland Geometry on Emission of Unburned Hydrocarbons from a Homogeneous Charge Compression Ignition (HCCI) Engine, SAE Paper 2001-01-1893, 2001. 23. Dingle, P. J. and Lai, M. D. (Ed.), Diesel Common Rail and Advanced Fuel Injection Systems, pp.29, SAE International, 2005 24. Kook, S., Bae, C., Miles, P. C., Choi, D, and Pickett, L. M. The Influence of Charge Dilution and Injection Timing on Low-Temperature Diesel Combustion and Emissions, SAE Paper 2005-01-3837, 2005.

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