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Title Study on Influence of Fuel Properties on Premixed Diesel Combustion

Author(s) 熊, 仟

Citation 北海道大学. 博士(工学) 甲第11448号

Issue Date 2014-03-25

DOI 10.14943/doctoral.k11448

Doc URL http://hdl.handle.net/2115/55503

Type theses (doctoral)

File Information Xiong_Qian.pdf

Hokkaido University Collection of Scholarly and Academic Papers : HUSCAP

https://eprints.lib.hokudai.ac.jp/dspace/about.en.jsp

STUDY ON INFLUENCE OF FUEL PROPERTIES

ON PREMIXED DIESEL CMBUSTION

A Dissertation submitted in partial fulfillment of the

requirement of the Doctor Degree (Engineering)

In Hokkaido University, JAPAN

2014

by

Qian Xiong

DISSERTATION ABSTRACT

Premixed diesel combustion, as a promising combustion concept to achieve low NOx and smoke

emissions as well as high thermal efficiency, is paid much attention. Sufficiently long ignition delay is

required for pre-mixture preparation to avoid over-rich mixture taking part in the combustion while the

maximum pressure rise rate is suppressed to a tolerance level. Therefore, the operational load range of

premixed diesel combustion with diesel fuel is limited at low and medium loads by the high pressure

rise rate when increasing the fuelling rate per cycle. The results in recent researches indicate,by late

injection strategy, with sufficiently low ignitability and high volatility fuel like gasoline, almost full

operational load and speed ranges can be established in diesel engines with higher thermal efficiency

and low emission levels. In this study, the influence of fuel properties including fuel ignitability,

compositions and volatility on premixed diesel combustion is systematically investigated by

experiments conducted on a single-cylinder, supercharged, four-stroke cycle, direct injection diesel

engine with low pressure loop cooled EGR and common-rail fuel injection systems to obtain the

optimal fuel properties for premixed diesel combustion.

Secondly, the stable premixed diesel operational range of intake oxygen concentrations for various

indicated mean effective pressures (IMEP) was constraints of knocking (maximum pressure rise rate

lower than 1.0 MPa/°CA), unstable combustion (the coefficient of variance (COV) in IMEP below 5%)

and the lower indicated thermal efficiency (higher than 40%).

Influence of ignitability on premixed diesel combustion is firstly discussed with four primary

reference fuels (PRFs) with octane numbers from 0 to 60. With increasing octane number, the

combustion phasing is retarded and higher intake oxygen concentration can be used resulting in higher

indicated thermal efficiencies due to the higher combustion efficiency and the larger maximum IMEP

of the premixed diesel combustion. Silent, low NOx, and smokeless operation with a high thermal

efficiency is possible up to 1.0 MPa IMEP using PRF with a research octane number of 40 when the

intake oxygen concentration was optimized corresponding to IMEP.

Next, influence of the fuel compositions on premixed diesel combustion is investigated with the

difference between PRF and normal heptane and toluene blend fuel (NTF) with the same research

octane number as the PRF. Compared with PRFs, the stable operational ranges and optimum intake

oxygen concentrations of premixed diesel combustion with NTFs with the same research octane

numbers shift to the higher intake oxygen concentration side, showing lower ignition characteristics

than PRFs. Slightly higher maximum IMEP can be achieved with NTF with premixed diesel

combustion, compared with PRF of the same research octane number.

Influence of fuel volatility on operational range and combustion characteristics of premixed diesel

combustion is investigated with normal heptane, PRF20 and an ordinary diesel fuel. The results show

the fuel volatility has little effect on the stable operational range and combustion characteristics of

premixed diesel combustion. However, the best indicated thermal efficiency with the ordinary diesel

fuel was lower than with PRF20 at the optimum intake oxygen concentration in a wide IMEP range,

mainly due to the higher unknown loss. The unknown loss defined here includes the cooling loss

(main component), the losses from unburned fuel entering the crank case and also the oxidized fuel

after the apparent heat release in the cylinder during the expansion and exhaust strokes as well as in

the exhaust manifold before the exhaust gas sampling. The deterioration of indicated thermal

efficiency with the ordinary diesel fuel here is mainly due to the increased quantity of fuel which

doesn’t contribute to the effective combustion, as the high distillation components in diesel fuel adhere

to the wall of the combustion chamber without evaporating.

Influence of fuel volatility on the thermal efficiency of premixed diesel combustion is investigated

with three ordinary diesel fuels with different distillation temperature distributions and PRF20 as a

high volatility fuel. With the premixed diesel combustion mode, regardless of injection timings, the

indicated thermal efficiencies with the ordinary diesel fuels are lower than with PRF20 due to the

higher unknown losses. With the conventional diesel combustion mode, the indicated thermal

efficiencies with the ordinary diesel fuels and PRF20 are similar. These results suggest that the

deterioration in the thermal efficiency in the premixed diesel combustion mode with the ordinary

diesel fuel is mainly due to the adhesion of high distillation temperature components on the walls of

the combustion chamber, and occurs mainly during the premixing period (between the end of fuel

injection and the ignition); the larger amount of unburned fuel enters the crank case and the fuel

oxidized after the apparent heat release in the cylinder do not contribute to the effective heat release.

Study on Influence of Fuel Properties on Premixed Diesel Combustion

i

TABLE OF CONTENTS

CHAPTER 1. INTRODUCTION

1.1 Overview of internal combustion engines .................................................................................. 1

1.1.1 Improvement in thermal efficiency ...................................................................................... 1

1.1.2 Reduction in emission levels ................................................................................................ 4

1.2 Progresses in premixed diesel combustion ................................................................................. 6

1.2.1 Principle of premixed diesel combustion ............................................................................. 6

1.2.2 Researches of premixed diesel combustion .......................................................................... 9

1.2.3 Researches on the effect of fuel properties on premixed diesel combustion ...................... 13

1.3 Motivation and objectives ........................................................................................................ 16

1.4 Layout of dissertation ............................................................................................................... 17

CHAPTER 2. EXPERIMENTAL APPARATUS AND PROCEDURE

2.1 Experimental apparatus ............................................................................................................ 19

2.2 Properties of the test fuels ........................................................................................................ 20

2.3 Fuel injection systems .............................................................................................................. 21

2.3.1 Nozzle specifications .......................................................................................................... 21

2.3.2 Common rail systems ......................................................................................................... 21

2.3.3 Injection strategy ................................................................................................................ 22

2.4 Operating conditions ................................................................................................................ 23

2.5 Measurement and analysis of indicator diagram ...................................................................... 25

2.5.1 Acquisition of in-cylinder pressure .................................................................................... 25

2.5.2 Heat release calculation methodology ................................................................................ 26

2.5.3 The degree of constant volume heat release ....................................................................... 28

2.6 Analysis of exhaust gas emissions ........................................................................................... 29

2.7 Other measurements ................................................................................................................. 30

CHAPTER 3. INFLUENCE OF FUEL IGNITABILITY ON OPERATIONAL RANGE AND

COMBUSTION CHARACTERISTICS OF PREMIXED DIESEL COMBUSTION

3.1 Chapter outline ......................................................................................................................... 33

3.2 The stable operational range and combustion characteristics of premixed diesel combustion

with diesel fuel ......................................................................................................................... 33

3.3 Effect of fuel ignitability on stable operational range of premixed diesel combustion ............ 37

3.4 Effect of fuel ignitability on combustion characteristics and emissions of premixed diesel

combustion ............................................................................................................................... 40

3.5 Chapter summary ..................................................................................................................... 44

CHAPTER 4. INFLUENCE OF FUEL COMPOSITIONS ON OPERATIONAL RANGE

AND COMBUSTION CHARACTERISTICS OF PREMIXED DIESEL COMBUSTION


ii

4.1 Chapter outline ......................................................................................................................... 46

4.2 Effect of fuel compositions on stable operational range of premixed diesel combustion ........ 46

4.3 Effect of fuel compositions on combustion characteristics and emissions on premixed diesel

combustion ............................................................................................................................... 49

4.4 Chapter summary ..................................................................................................................... 53

CHAPTER 5. INFLUENCE OF FUEL VOLATILITY ON OPERATIONAL RANGE AND

COMBUSTION CHARACTERISTICS OF PREMIXED DIESEL COMBUSTION

5.1 Chapter outline ......................................................................................................................... 55

5.2 Effect of fuel volatility on stable operational range of premixed diesel combustion ............... 55

5.3 Effect of fuel volatility on combustion characteristics and emissions on premixed diesel

combustion ............................................................................................................................... 56

5.4 Chapter summary ..................................................................................................................... 61

CHAPTER 6. INFLUENCE OF FUEL DISTILLATION TEMPERATURE ON THERMAL

EFFICIENCY OF PREMIXED DIESEL COMBUSTION

6.1 Chapter outline ......................................................................................................................... 62

6.2 Simulation of the evaporation processes of one fuel droplet with n-heptane and n-tridecane in

the test engine condition ........................................................................................................... 62

6.3 Effect of fuel distillation temperature on thermal efficiency of premixed diesel combustion

with different injection timings ................................................................................................ 64

6.4 Influence of fuel volatility on thermal efficiencies in premixed diesel combustion and in

conventional diesel combustion ............................................................................................... 70


with different coolant temperatures .......................................................................................... 72

6.6 Effect of fuel volatility on thermal efficiency of premixed diesel combustion with different

intake gas temperatures ............................................................................................................ 76

6.7 Chapter summary ..................................................................................................................... 79

CHAPTER 7. SUMMARY .............................................................................................................. 81

REFERENCES........................ ............................................................................................................ 83


iii

LIST OF FIGURES

Figure 1-1 Forecast of future world oil production ................................................................................ 1

Figure 1-2 The expectation of energy-related carbon emissions predicted by the International Energy

Agency, 2013........................................................................................................................ 2

Figure 1-3 History of the carbon emissions sources ............................................................................... 2

Figure 1-4 The predicted number of vehicles per thould persons at 2010 and 2035 in some regions of

the world ............................................................................................................................... 3

Figure 1-5 History of the trace of global temperature and carbon dioxide ............................................ 3

Figure 1-6 The histories of emission regulations in NOX and PM on diesel passenger cars in Europe,

the USA and Japan ............................................................................................................... 4

Figure 1-7 Configuration of an emission reduction technology ............................................................. 5

Figure 1-8 The trace of the sulfur content in gasoline and diesel fuels in Europe, the USA and Japan . 6

Figure 1-9 The operating principles of spark ignition gasoline engine, traditional compression ignition

diesel engine, and premixed charge compression ignition engine ........................................ 7

Figure 1-10 The plots of the in-cylinder pressure and the rate of heat release in conventional diesel

combustion mode and premixed diesel combustion mode ................................................ 8

Figure 1-11 PCCI combustion process in comparison to other methods on -T map ........................... 9

Figure 1-12 Operation map with UNIBUS (LAND CRUISER PRADO) ........................................... 10

Figure 1-13 MK combustion range of the first generation (YD25DDT) and target of the second

generation ........................................................................................................................ 13

Figure 2-1 Experimental arrangements ................................................................................................ 19

Figure 2-2 The common rail fuel injection system ............................................................................... 22

Figure 2-3 Positions of the piston and fuel spray at different fuel injection timings under the premixed

diesel combustion mode ..................................................................................................... 23

Figure 2-4 In-cylinder pressure acquisition system .............................................................................. 26

Figure 2-5 Segmental Otto cycle .......................................................................................................... 29

Figure 2-6 Dynamometer ..................................................................................................................... 30

Figure 2-7 Fuel flowmeter .................................................................................................................... 30

Figure 2-8 Surge tank ........................................................................................................................... 31

Figure 2-9 Digital manometer .............................................................................................................. 31

Figure 2-10 Portable oxygen tester (POT-101: SHIMAZU) ................................................................ 31

Figure 3-1 Practical premixed diesel operational range of intake oxygen concentrations, O2in, and

excess air ratios with JIS No.2 diesel fuel ......................................................................... 34

Figure 3-2 Optimum intake oxygen concentrations, O2in, and excess air ratios for the best indicated

thermal efficiencies with JIS No.2 diesel fuel ................................................................... 34


iv

Figure 3-3 Thermal efficiency related parameters with diesel fuel at the optimum intake oxygen

concentrations ................................................................................................................... 36

Figure 3-4 Exhaust gas emissions and the maximum rates of pressure rise, dp/dmax with diesel fuel at

the optimum intake oxygen concentrations ........................................................................ 37

Figure 3-5 Stable premixed diesel operational range of intake oxygen concentrations, O2in with PRF0

and PRF60 .......................................................................................................................... 38

Figure 3-6 In-cylinder pressure and the rate of heat release (ROHR) with PRF0 at O2in = 16.2% and

PRF60 at O2in = 16.5% (IMEP: 0.4 MPa) .......................................................................... 39


PRF60 at O2in = 17% (IMEP: 0.6 MPa) ............................................................................. 39

Figure 3-8 The optimum intake oxygen concentrations, O2in and the excess air ratios for the best

indicated thermal efficiencies with PRF0 and PRF40 ....................................................... 40

Figure 3-9 Thermal efficiency related parameters with PRF0 and PRF40 at the optimum intake

oxygen concentrations ...................................................................................................... 41

Figure 3-10 Combustion related parameters with PRF0 and PRF40 at the optimum intake oxygen

concentrations .................................................................................................................. 43

Figure 3-11 Exhaust gas emissions and the maximum rates of pressure rise with PRF0 and PRF40 at

the optimum intake oxygen concentrations ...................................................................... 44


and NTF17 .......................................................................................................................... 47


and NTF34 .......................................................................................................................... 47


and NTF50 .......................................................................................................................... 48

Figure 4-4 Effect of octane number on the maximum IMEP and the intake oxygen concentration at the

maximum IMEP for PRF and NTF ..................................................................................... 48

Figure 4-5 In-cylinder pressure and the rate of heat release (ROHR) with PRF40 and NTF34 under the

same intake oxygen concentrations of 14.4% (IMEP: 0.6 MPa) ........................................ 49

Figure 4-6 The optimum intake oxygen concentration, O2in and the excess air ratio for the best

indicated thermal efficiency with PRF40 and NTF34 ...................................................... 50

Figure 4-7 Thermal efficiency related parameters with PRF40 and NTF34 at the optimum intake

oxygen concentrations ....................................................................................................... 51

Figure 4-8 Combustion related parameters with PRF40 and NTF34 at the optimum intake oxygen

concentrations .................................................................................................................... 52

Figure 4-9 Exhaust gas emissions and the maximum rates of pressure rise, dp/dmax with PRF40 and

NTF34 at the optimum intake oxygen concentrations ....................................................... 53


v

Figure 5-1 Stable premixed diesel operational range of intake oxygen concentrations, O2in

with

ordinary diesel fuel (JIS No.2 diesel fuel) and normal heptane for various IMEP ........... 55

Figure 5-2 In-cylinder pressure and the rate of heat release (ROHR) with diesel fuel and normal

heptane at O2in

= 11.4% (IMEP: 0.6 MPa) ........................................................................ 56

Figure 5-3 Exhaust gas emissions with diesel fuel and normal heptane for various IMEP at the

knocking limits .................................................................................................................. 57

Figure 5-4 The optimum intake oxygen concentration, O2in

, for the best thermal efficiency with JIS

No.2 diesel fuel and PRF20 ............................................................................................... 57

Figure 5-5 The thermal efficiency related parameters with JIS No.2 diesel fuel and PRF20 at the

optimum intake oxygen concentration conditions ............................................................. 58

Figure 5-6 In-cylinder pressure and the rate of heat release (ROHR) with JIS No.2 diesel fuel and

PRF20 at the three maximum rates of pressure rise (IMEP: 0.6 MPa) .............................. 59

Figure 5-7 Intake oxygen concentrations, O2in, and thermal efficiency related parameters with JIS

No.2 diesel fuel and PRF20 for the maximum rate of pressure rise, dp/dmax (IMEP:

0.6MPa) ............................................................................................................................. 60

Figure 6-1 Spray tip penetrations of n-heptane in the conditions of the test engine at 40ºCA BTDC

(2.0 MPa and 650 K) and 25ºCA BTDC (3.5 MPa and 750 K) ......................................... 63

Figure 6-2 Changes in the residual mass fraction of a fuel droplet at (a) 40ºCA BTDC and (b) 25ºCA

BTDC (Initial droplet diameter: 10 um) ............................................................................. 64

Figure 6-3 Thermal efficiency related parameters for special No. 1 diesel fuel, special No. 3 diesel

fuel, and PRF20 for three fuel injection timings under the premixed diesel combustion

mode (Coolant temperature: 80ºC, Intake gas temperature: 40ºC) .................................... 66

Figure 6-4 The average in-cylinder temperatures, in-cylinder pressures and the rates of heat release

(ROHR) with special No. 1 diesel fuel, special No. 3 diesel fuel, and PRF20 for three fuel

injection timings under the premixed diesel combustion mode (Coolant temperature: 80ºC,

Intake gas temperature: 40ºC) ............................................................................................ 67

Figure 6-5 The rate of cooling loss per degree crank angle, dQw/d, with special No.1 diesel fuel,

special No.3 diesel fuel and PRF20 for three fuel injection timings under the premixed

diesel combustion mode (Coolant temperature: 80ºC) ...................................................... 68

Figure 6-6 Combustion related parameters with special No.1 diesel fuel, special No.3 diesel fuel and

PRF20 for three fuel injection timings under the premixed diesel combustion mode

(Coolant temperature: 80ºC) .............................................................................................. 69

Figure 6-7 Exhaust gas emissions with special No.1 diesel fuel, special No.3 diesel fuel, and PRF20

for three fuel injection timings under the premixed diesel combustion mode (Coolant

temperature: 80ºC) ............................................................................................................. 70


vi


(ROHR) with special No.1 diesel fuel and PRF20 at 0.7 IMEP for the conventional diesel

combustion mode ............................................................................................................... 71

Figure 6-9 Energy balances with special No.1 diesel fuel and PRF20 for the conventional diesel

combustion premixed diesel combustion mode and premixed diesel combustion mode

(inj : 10ºCA BTDC) .......................................................................................................... 72

Figure 6-10 Thermal efficiency related parameters with special No.1 diesel fuel and PRF20 for three

fuel injection timings at coolant temperatures of 80ºC and 40ºC under the premixed

diesel combustion mode ................................................................................................... 73


(ROHR) with special No.1 diesel fuel and PRF20 for two injection timings at coolant

temperatures of 80ºC and 40ºC under the premixed diesel combustion mode................. 75

Figure 6-12 Thermal efficiency related parameters for DF1 and PRF20 for three fuel injection timings

at two intake gas temperatures under the premixed diesel combustion mode (Coolant

temperature: 80ºC) ............................................................................................................ 77


(ROHR) With No.1 diesel fuel and PRF20 for two injection timings under the premixed

diesel combustion mode (Coolant temperature: 80ºC) ..................................................... 78


vii

LIST OF TABLES

Table 2-1 Test engine specifications .................................................................................................... 19

Table 2-2 Properties of the tested fuels for the influence of fuel properties on the operational range

and combustion characteristics of premixed diesel combustion in chapter 3, chapter 4, and

chapter 5 .............................................................................................................................. 20

Table 2-3 Properties of the tested fuels for the effect of fuel volatility on the thermal efficiency of the

premixed diesel combustion in chapter 6 ............................................................................. 20

Table 2-4 Nozzle specifications ............................................................................................................ 21

Table 2-5 The operating conditions for fuel properties on operational range and combustion

characteristics of premixed diesel combustion in chapter 3, chapter 4, and chapter 5 ...... 23

Table 2-6 The operating conditions for fuel distillation temperature on thermal efficiency of premixed

diesel combustion in chapter 6 ............................................................................................. 25

Table 2-7 Constant of Tanishita’s specific heat equation ................................................................... 27


CHAPTER 1. INTRODUCTION


1

1.1 Overview of internal combustion engines

As the main propulsion of automobiles for about 140 years, internal combustion engines have been

invented dating back to 1876 when N. Otto first developed the spark-ignition engine and 1892 when R.

Diesel developed the compression-ignition engine. Although many modifications have been made to

improve the performance of internal combustion engines, there are still two challenges for future

internal combustion engines.

Improvement in thermal efficiency

Reduction in emission level

1.1.1 Improvement in thermal efficiency

While that of the first production four-stroke gasoline engines developed by N. Otto was only 16%,

the thermal efficiency of modern internal combustion engines has been increased to higher than 35%.

However, it should be further improved in future. The necessity in improving the thermal efficiency of

internal combustion engines will be explained as follows.

(a) Depletion of petroleum resources

Figure 1-1 Forecast of future world oil production [1]

After the innovation of internal combustion engines, the oil consumption over the world was gradually

increased. As Figure 1-1 released from the Association for the Study of Peak Oil at Uppsala shows, the

oil production increased largely from 1950, and until 2015 it will reach the peak value. After that, the

oil production will gradually decrease which suggest the depletion of petroleum resources will occur.


2

(b) Increase in fuel consumption and global warming

As the economic development of the emerging markets over the world, the amounts of fuel

consumption are increasing rapidly. Referencing the report of the global energy use released by the

International Energy Agency as shown in Figure 1-2, the gross amount of carbon emissions has been

increasing significantly for 40 years. Although across the European Union and Japan, the amount of

carbon emissions will decline based on current promise, those will significantly increase in China and

India in the following 30 years.

Figure 1-2 The expectation of energy-related carbon emissions predicted by the International Energy

Agency, 2013 [2]

Figure 1-3 History of the carbon emissions sources [3]

In the construction of the global fossil carbon emissions, the petroleum is the main source of the

carbon emissions, resulting in half of the amount, as shown in Figure 1-3.


3

Meanwhile, the number of the passenger cars, which use the gasoline or diesel fuel obtained from the

refinery of petroleum as the main energy source, is still gradually increasing. Referencing the world

energy outlook released by International Energy Agency as Figure 1-4, the vehicles per thousand

persons will increase in all the areas shown here and particularly, the growth rates of that in the

developing countries, for example, China, India and Middle East are very large in the following 25

years.

Figure 1-4 The predicted number of vehicles per thould persons at 2010 and 2035 in some regions of

the world [4]

Figure 1-5 History of the trace of global temperature and carbon dioxide [5]

Global warming is paid more attention recently, and is the rise in the average temperature of earth's

atmosphere and oceans since the late 19th century and its projected continuation. As the analysis

results from the researchers in Carbon Dioxide Information Analysis Center of United States shown in

Figure 1-5, since the early 20th century, average surface temperature on earth has increased by about


4

0.8ºC (1.4ºF), with about two-thirds of the increase occurring since 1980. Warming of the climate

system is unequivocal, and scientists are 95-100% certain that it is primarily caused by increasing

concentrations of greenhouse gases (mainly CO2, carbon dioxide) produced by human activities such

as the burning of fossil fuels and deforestation. These findings are recognized by the national science

academies of all major industrialized nations. In terms of engines, CO2 is their main exhaust gas

component.

1.1.2 Reduction in emission levels

Besides CO2, the other noxious emissions including carbon monoxide (CO), unburned hydrocarbons

(UHC), oxides of nitrogen (NOX) and particulate matters (PM) are also emitted to atmosphere from the

automobiles. These emissions should be significantly reduced in future.

(a) Emission regulations

(i) Oxides of nitrogen (NOX) (ii) Particulate matters (PM)

Figure 1-6 The histories of emission regulations in NOX and PM on diesel passenger cars in Europe,

the USA and Japan [7]

As the automotive ownership in the world gradually increased, the automotive air-pollution problem

became apparent in the 1940s in the Los Angeles basin. The smog problem was firstly demonstrated

by Prof. A. J. Haagen-Smit in 1952, which resulted from reactions between NOX and hydrocarbon

compounds in the presence of sunlight [6]. Gasoline and diesel engines are the significant sources of

UHC, CO, and smoke, as well as NOX. As a result of these developments, emission standards for


5

automobiles were introduced by the government first in California, then nationwide in the United

States, starting in the early 1960s. After that, emission standards in the main market of automobile

over the world, and for other engine applications, have followed, and been gradually stringent. The

histories of emission regulations of NOX and particulate matters (PM) on diesel passenger cars in

Europe, the USA and Japan during 20 years are shown as Figure 1-6. At present, the levels of NOX

and PM in these regions are limited to below 0.1 g-CO2/km and 0.01 g-CO2/km respectively.

(b) Emission control technology

To reduce the noxious exhaust gas into atmosphere and meet the stringent emission regulations, a lot

of emission control technologies have been innovated and equipped at the downstream of the exhaust

manifold of the engines. Figure 1-7 shows one typical configuration of aftertreatment in diesel

engines.

Figure 1-7 Configuration of an emission reduction technology [8]

The work process of these instruments is as follows:

① PM is captured in the diesel particulate filter (DPF);

THC (Total hydrocarbons) and CO are converted to harmless water vapor and CO2;

② Aqueous urea solution is injected into hot exhaust gas stream;

The heat of exhaust converts the urea gas stream solution to ammonia gas;

③ NOX reacts with ammonia gas (NH3) to form nitrogen (N2) and water vapor using Selective

Catalytic Reduction (SCR) catalyst;

④ Residual CO and THC are removed with the oxidation catalyst;

⑤ Clean exhaust gases exit the stack.

There are many kinds of diesel particulate filters [9], including cordierite wall flow filters, silicon

carbide wall flow filters, ceramic fiber filters and metal fiber flow through filters. Some filters are


6

single-use, intended for disposal and replacement once full of accumulated ash. Others are designed to

burn off the accumulated particulate either passively through the use of a catalyst or by active means

such as a fuel burner which heats the filter to soot combustion temperatures, namely as “filter

regeneration”. Ceramic materials, including silicon carbide (SiC), ceramic fibers and metal fibers,

which are used to manufacture the filters, are expensive. To get high exhaust temperature for filter

regeneration process, more fuel is required to be injected in the cylinder.

In SCR catalysts system, another aqueous urea supply system is necessary to provide NH3 to induce a

chemical reaction with NOX, resulting in relative increase in fuel consumption. In terms of fuel, these

devices should be always used combining with diesel fuels of substantially lowered sulfur content,

resulting in higher cost of oil production when refining the crude oil. The trends of sulfur content in

the gasoline and diesel fuels at Europe, the USA, and Japan are shown as Figure 1-8.

(i) Gasoline (ii) Diesel fuel

Figure 1-8 The trace of the sulfur content in gasoline and diesel fuels in Europe, the USA and Japan

[10]

1.2 Progresses in premixed diesel combustion

Premixed Diesel Combustion (or Premixed Charge Compression Ignition, PCCI) was paid much

attention due to simultaneous reductions of NOX and smoke in diesel engines while high thermal

efficiency like that with conventional diesel combustion can be maintained. With this combustion

system, the expensive aftertreatment may be unnecessary in future to meet the emission regulations.

1.2.1 Principle of premixed diesel combustion

By reviewing the operating principles of spark ignition gasoline engines, conventional compression

ignition diesel engines, and premixed charge compression ignition engines, the characters of premixed

diesel combustion will be described in detail.


7

As Figure 1-9 shows, in spark ignition gasoline engines, the fuel is inducted into the cylinder either by

port injection in the intake manifold or by in-cylinder direct injection. Due to the relatively high

volatility of gasoline fuel, homogeneous charge can be obtained and ignition is initiated by a spark

plug. The NOX emissions often form in the high-temperature burned gas regions, as shown in Figure

1-9. In conventional diesel combustion, fuel is directly injected into the cylinder near the top dead

center (TDC) in the compression stroke just before combustion starts, and the in-cylinder fuel

distribution is very non-uniform. The ignition in conventional diesel combustion is initiated in much

higher temperature atmosphere than the ignition point. The process of conventional diesel combustion

can be divided into two phases: the “premixed” phase and “mixing-controlled” phase. The NOX

emissions often form in the high-temperature burned gas regions (higher than 2000K, particularly in

the close-to-stoichiometric regions) where the equivalence ratio is still sufficiently low (below two)

during two phases. Soot always forms in the rich unburned–fuel-containing core of the fuel sprays,

within the flame region, where the fuel vapor is heated by mixing with hot burned gases just after

ignition. In premixed diesel combustion (or PCCI), the liquid fuel is injected in the cylinder early

before TDC and is premixed sufficiently with air before combustion starts. Here, the spontaneous

ignitions at many regions are initiated when the local in-cylinder temperatures increase and reach the

ignition point during the compression stroke.

Figure 1-9 The operating principles of spark ignition gasoline engines, conventional compression

ignition diesel engines, and premixed charge compression ignition engines [11]


8

To show the difference of combustion process between conventional diesel combustion and premixed

diesel combustion further, the plots of the in-cylinder pressure and the rate of heat release with these

two combustion modes are shown as Figure 1-10. There is a long premixing duration from the end of

fuel injection to ignition in premixed diesel combustion mode, while the ignition occurs before the end

of fuel injection in conventional diesel combustion mode. From the (equivalence ratio) - T (local

adiabatic flame temperature) map as shown in Figure 1-11, the reason why simultaneous reductions of

NOX and smoke can be realized in premixed diesel combustion mode is as follows:

Due to the long ignition delay, the uniform premixed mixture can be formed and the region of

over-rich mixture can be removed before the ignition, inhibiting the formation of soot. Meanwhile,

ultra-high rate of EGR (Exhaust Gas Recirculation) was always applied in premixed diesel combustion

to achieve long ignition delay and decrease the local adiabatic flame temperature (<1900K) region and

local oxygen concentration region, suppressing the formation of NOX.

Figure 1-10 The plots of the in-cylinder pressure and the rate of heat release in conventional diesel

combustion mode and premixed diesel combustion mode


9

Figure 1-11 PCCI combustion process in comparison to other methods on -T map [12]

LTC: Low temperature combustion

HCCI: Homogenous charge compression ignition

PCCI: Premixed charge compression ignition

1.2.2 Researches of premixed diesel combustion

So far, many strategies to establish premixed diesel combustion in practical use have been attempted.

Three representative strategies will be described in detail as follows.

(a) Early in-cylinder injection

UNIBUS (Uniform Bulky Combustion System)

To obtain premixed fuel and air mixture with diesel fuel and realize the premixed diesel combustion in

DI diesel engines as the HCCI concept in gasoline engines, Yanagihara [13-14] from Toyota motor

company firstly used a single early direct injection (around 50ºCA BTDC), short injection duration

and low fuel injection pressure with a fuel spray guided injector to reduce the spray penetration and

minimize the fuel wall wetting in one DI diesel engine. Very low NOX and PM emissions combustion

could be achieved but only operate in a limited load range, while the levels of UHC and CO emissions

were very high (about 5000ppm), comparing with that in conventional diesel combustion mode. This

was mainly due to advanced ignition timing by early direct injection resulting in knocking and the

ignition timing was mainly controlled by the low temperature reaction.

To extend the possible operational load range and decrease the UHC and CO emissions, a

dual-injection strategy with common rail injection system was developed by Hasegawa et al. [15],

with 50% of the fuel introduced in a second injection at 13ºCA ATDC. The early injection was used


10

for fuel diffusion and to advance the decomposition of diesel fuel to lower hydrocarbons in low

temperature oxidation process. The second injection was used as an ignition trigger for all the injected

fuel. As the results showed, the ignition timing could be controlled by the second injection when the

early injection was maintaining low temperature oxidation, and the UHC and CO emissions could be

reduced from 5000 ppm to 2000 ppm and about 8000 ppm to 2000 ppm, respectively. However, the

operational range of premixed diesel combustion with this strategy was also limited within half loads

as shown in Figure 1-12. The concept of UNIBUS has already been applied to the production engine

(1KD-FTV, 3 liter-4cylinder) in August 2000 in the Japanese market.

Figure 1-12 Operation map with UNIBUS (LAND CRUISER PRADO) [15]

PREDIC (Premixed Lean Diesel Combustion)

Takeda et al. [16] and Nakagome et al. [17] from the New ACE Institute in Japan also used two stage

direct fuel injections and high fuel injection pressure (150 MPa and 250 MPa) to realize the premixed

diesel combustion mode with diesel fuel. Three injectors were used (one at the

center0.17mm6-155o or 0.08mm8-155o, and two on the sides, 0.17mm2) and the injection

timing and quantity of each injector independently. The double injectors mounted on either side of the

cylinder were positioned so that the fuel sprays would collide in the middle of the chamber to limit

wall-wetting with one very early injection (80oCA BTDC) and produce the lean premixed charge,

whilst a center-mounted injector was used to inject additional fuel relatively near TDC (40oCA BTDC).

By this injection strategy, the engine could be operated with PREDIC mode. The extremely low NOX

emissions are due to a very lean fuel-air ratio and gentle heat release pattern. However, the UHC and

CO emissions were still very high with deterioration of combustion efficiency, and there was still

lacking a direct control of ignition timing. And the results showed higher fuel consumption and

combustion speed, which limited the engine operating range. Several species of inert gases were


11

mixed with the intake air and with CO2 as the diluent gas, namely EGR, showed the best promise in

terms of ignition timing control.

To extend the operational range in high load side, this early-DI work was expanded to include a

second stage injection near TDC (from 2oCA BTDC to 30oCA ATDC) with the center injector [18].

Approximately half of the fuel was injected in each stage, which was near the maximum amount of

premixed fuel without knocking and higher operation load range was possible to reach with this

injection strategy. As the combustion with second stage injection was similar to conventional diesel

combustion, NOX and smoke emissions were much higher with this strategy than those obtained with

pure premixed diesel combustion. However, the NOX emissions were reduced to about half of that

with conventional diesel combustion without a significant penalty of fuel consumption, and smoke

emissions also decreased. The HC and CO emissions were still very high. At last, as described as

before, the injection strategy of PREDIC system was much complicated. Further researches show that

THC and CO emissions could be decreased by adopting a pintle type injection nozzle, or a reduced

top-land-crevice piston [19]. Fuel consumption was also improved by applying EGR or the addition of

an oxygenated component to the diesel fuel. As a result, the operating region could be extended to

higher load conditions by MULtiple stage DIesel Combustion (MULDIC), in which the first stage

combustion corresponds to PREDIC.

(b) Narrow angle direct injection

The use of a narrow fuel injector nozzle spray cone angle is an aim to reduce the fuel wetting on the

piston head surface and cylinder wall at the early injection timings. There are several reports to show

low NOX and smoke premixed diesel combustion with this strategy while maintaining a relatively high

thermal efficiency [20-23].

The IFP has developed a combustion system which is able to reach near smokeless and NOX emissions

while maintaining performance standards of the DI diesel engines [20]. This dual mode engine

application called NADITM (narrow angle direct injection) applies premixed diesel combustion at low

and medium loads and switches to conventional diesel combustion to reach high and full loads

requirements. The utilization of one narrow spray cone angle injector (lower than 100º) can reduce the

wall wetting problem and avoid an out of piston bowl injection when the fuel was injected at an early

timing for premixed diesel combustion. In the premixed diesel combustion mode, a lot of development

studies have been done on the control of combustion with the use of high EGR rate, variable effective

compression ratio thanks to valve timing variations, intake gas temperature and injection timing.

However, the results in this study show the operational load range of premixed diesel combustion is

still to be extended and the CO and THC emissions are too high. The transition strategies between the

two combustion modes should be further developed.


12

In the research of Kim and Lee [24] the effect of narrow fuel injector nozzle spray cone angle on

premixed diesel combustion was investigated with one injector nozzle with the spray cone angle of

156º at a compression ratio of 17.8:1 and another injector nozzle with 60º at 15:1. The results showed

that in the case of the conventional diesel engine, the IMEP was decreased rapidly as the injection

timings were advanced beyond 20ºCA BTDC and thermal efficiencies were also deteriorated. In the

case of advanced injection timing between 30ºCA and 50ºCA before the TDC, the IMEP was

approximately half of that attained under the conventional diesel combustion with the injection timing

near the TDC, a too early injection leads to poor fuel evaporation and piston bowl spray targeting

issues. At that condition, the pre-mixture is formed outside the combustion chamber, resulting in

deteriorated combustion efficiency and the increase in negative work during compression stroke with

too early combustion. In contrast, in the case of a narrow spray angle configuration, the ISFC

indicated a modest decrease in the IMEP even if the injection timing was advanced to 50~60ºCA

BTDC. This reveals that the narrow angle concept was effective in maintaining the low ISFC and high

IMEP when the fuel was injected at an early timing for premixed diesel combustion. With a second

injection near the TDC with a narrow spray cone angle injector, higher IMEP can be attained in

premixed diesel combustion.

Lechner et al. [25] also investigated the effect of an injector with a narrow spray cone angle and small

diameter holes on premixed diesel combustion. The PM emissions are decreased due to the longer

premixing duration and the short spray penetration with split fuel injections but the fuel consumption

was deteriorated. André et al. [26] proposed an optimization of the injection strategy including the

injection timing and the injection pressure maximizing the fuel quantity without cylinder wall wetting

and tested by experiments conducting on a low swirl, low compression ratio engine with a narrow fuel

spray angle injector. In this case, the operational range of the load was limited to 0.6 MPa IMEP with

only very early direct injections.

(c) Late in-cylinder injection

The MK (modulated kinetics) combustion system developed by the Nissan Motor Company is one of

the most successful premixed diesel combustion systems with in-cylinder late direct injection near the

top dead center (TDC) for achieving diesel-fuelled premixed diesel combustion [27-33]. The principle

of this combustion process was described by Kawashima et al. [28] and Kimura et al. [32-33].

To achieve the diluted homogeneous mixture required for premixed diesel combustion before the

ignition, sufficiently long ignition delay is obtained with retarding the injection timing from 7ºCA

BTDC to 3ºCA ATDC and high levels of hot EGR in this combustion system. Further rapid mixing is

achieved with combining high swirl with toroidal combustion-bowl geometry. Comparing with early

direct injection premixed diesel combustion systems, a distinct advantage of MK combustion system is


13

that the combustion phasing could be controlled by varying injection timing. In the MK mode, NOX

emissions were significantly reduced to about 50 ppm and the pressure rise rate was also significantly

reduced due to the retarded combustion phasing. However, the operating range for the first generation

MK system was limited to about one-third of peak torque and half-speed.

To extend the operation to higher loads and speeds for MK system, several modifications were

attained. A high injection pressure (130 MPa) with common-rail fuel system was applied to shorten

the fuel injection duration at all speeds. To achieve longer ignition delay, the compression ratio of the

engine was reduced to 16 and cooled EGR was used to reduce the intake temperature. For minimize

the quantity of liquid fuel impinging on the piston bowl wall, the diameter of the piston bowl was

increased from 47 mm to 56 mm. Due to these modifications, THC emission could be significantly

reduced under cold-engine conditions. Therefore, the second-generation MK system could be used

over the entire range of everyday driving, and the operating range extended to about half load and

three-quarters speed as shown in Figure 1-13. NOX emissions are reduced by 98% compared with

conventional operation without EGR, and the increases in fuel consumption and smoke formation

were suppressed.

Figure 1-13 MK combustion range of the first generation (YD25DDT) and target of the second

generation [33]

1.2.3 Researches on the effect of fuel properties on premixed diesel combustion

As mentioned in the previous researches reviewed in the last section, ultra-high EGR is an effective

way to obtain sufficiently long ignition delay and the fuel and air can be premixed well before the

combustion starts, resulting in simultaneous reductions of NOX and smoke emissions in premixed

diesel combustion. This strategy is always used combined with varying the injection timings and swirl


14

ratios. However, the utilization of ultra-high EGR also limits the extension of operational load range

with premixed diesel combustion to higher load.

In terms of fuels, fuel properties including ignitability and volatility have significant influence on the

fuel-air mixing and combustion processes [34-51]. The ignitability is directly related to the ignition

delay, which is the key to extend the operational range of premixed diesel combustion to higher loads.

The fuel distillation temperature and viscosity have important effect on the fuel evaporation process

after injection and further the mixing process of fuel and air. Some reviews about fuel properties on

premixed diesel combustion are as follow.

It is well known that the ignition quality of a diesel fuel is defined as a cetane number and the effect of

cetane number on premixed diesel combustion has been examined by many researchers. Li et al. [34]

found that with decreasing fuel cetane number, significant reduction of smoke emission could be

achieved due to the longer ignition delay. Furthermore, the operational range of smokeless premixed

diesel combustion could be significantly extended to higher load (0.58 MPa IMEP) with the fuel at

cetane number of 40.

Further investigation was carried out on the characteristics of premixed diesel combustion with fuels

of different cetane numbers at various fuel-air mixing conditions [37]. The results showed that when

the combustion temperature was sufficiently low with large quantities of cooled EGR, the NOX is

suppressed to near zero levels and if the ignition dwell (from the end of fuel injection to the onset of

ignition) is longer than about 4ºCA, the smokeless combustion could be achieved regardless of fuel

cetane numbers and fuel injection timings in the tested conditions. Recent research by Li et al. [39]

described that the fuel ignitability on premixed diesel combustion with the fuels of different predicted

cetane numbers (PCN) based on fuel molecular structure analysis could be fitted to the ignition delays.

With fixed injection timings and intake oxygen concentration, the operational load range could be

extended significantly to higher loads only when fuel PCN was lower than 40. Ickes et al. [40] also

investigated the effect of fuel cetane number on a premixed diesel combustion mode with cetane

numbers ranging from 42 to 53. The results showed that the operating range of premixed diesel

combustion was constrained by the changes in cetane number. However, in terms of combustion

phasing, the operation limit was independent on fuel cetane number.

Based on the strategy of Nissan MK style combustion, Kalghatgi et al. [41-42] studied the effect of

fuel auto-ignition quality, using four fuels ranging from diesel to gasoline, on premixed diesel

combustion at two intake gas pressures (1.5 bar and 2.0 bar) and different EGR levels. A

single-cylinder heavy duty diesel engine with a low compression ratio of 14 was used in this

experiment at an engine speed of 1200 rpm. With gasoline, the tested engine could be easily run in

premixed diesel combustion mode at a single injection near TDC. Moreover for any condition,


15

gasoline has a significantly longer ignition delay for the same combustion phasing, resulting in very

much lower NOX and smoke emissions for a given load compared to diesel fuels. Using gasoline, the

operational load range could be extended to an IMEP of 14.86 bar at 2.0 bar inlet pressure and 25%

EGR condition while high thermal efficiency can be maintained. These results suggested the

possibility of using the auto-ignition resistance of fuels like gasoline to attain high IMEP in low NOX,

low smoke combustion systems similar to the Nissan MK system which aim to promote pre-mixing

before ignition. To further understand the effects of fuel octane number on this combustion regime,

which was named as Partially Premixed Combustion (PPC) later, Hildingsson et al. [43-44]

investigated the PPC operation of a single-cylinder light duty compression ignition engine with four

different fuels of different research octane numbers (RON) of 91, 84, 78, and 72 in the gasoline

boiling range, and gave the comparison with that running on a diesel fuel. There was a NOX advantage

for the higher RON fuels (91 and 84) at the lower load (4 bar IMEP/1200 rpm). At the higher load (10

bar IMEP / 2000 and 3000 rpm), NOX levels can be reduced by increasing EGR for all gasolines while

maintaining much lower smoke levels compared to the diesel fuel. In this studied conditions, the

results suggested that the optimum RON range might be between 75 and 85 for PPC.

Ra et al. [45] presented an investigation of high speed DI compression ignition engine combustion

fueled with gasoline in the premixed diesel combustion regime. Operation ranges of the premixed

diesel combustion in this study are constraints of combustion efficiency (higher than 85%), the

maximum pressure rise rate (lower than 0.7 MPa/ºCA) and smoke and NOX emissions (lower than 1.0

g/kg-fuel) were identified as functions of injection timings, the EGR rate and the fuel split ratio of two

stage injections. The results showed that, due to the high octane number of gasoline combined with

utilization of EGR for suppressing in-cylinder combustion temperature, both NOX and smoke

emissions could meet the given constraint while maintaining low fuel consumption (gross ISFC at

about 180 g/kW·h). However, it was not possible to get comparable performance with diesel fuels

with low smoke at the same operating conditions. At last, maps of operable conditions were generated

that allow extension of low-emission engine concepts to full load operation in high speed DI diesel

engine operation with high thermal efficiency.

After review of ignitability, the researches about the effect of fuel volatility on premixed diesel

combustion should be described. As discussed by Epping et al. [46], the high volatility fuel has an

advantage to form a homogeneous charge. Sakai et al. [47] found that with early direct injection

timings (early before 60ºCA BTDC), wall wetting could be eliminated and the thermal efficiency in

premixed diesel combustion mode could be improved with high volatility fuel which distillation

temperature is lower than the in-cylinder gas temperature early in compression stroke. Li et al. [35]

also investigated the effect of fuel distillation temperature on premixed diesel combustion with fuels of

two cetane numbers (40 and 60) with different distillation temperatures. The results showed that the


16

distillation temperature had almost no influence on premixed diesel combustion while the injection

timing was adjusted to fix the ignition timing at TDC. Many other researches’ results shows similar

conclusion about the effect of fuel volatility on premixed diesel combustion [48-51].

1.3 Motivation and objectives

After the detailed review of pioneering researches, the challenges to premixed diesel combustion are

summarized as follow.

1. Extension of the operational load range to full load range, particularly high load side, as

conventional diesel combustion with diesel engines

2. The strategy for direct control of ignition timing

3. The decreases of UHC and CO emissions

4. Minimizing the wall wetting and oil dilution to improve the thermal efficiency

As a promising way to realize simultaneous reductions of NOX and smoke emissions and meet the

current pollutant emission regulations without expensive aftertreatment systems, premixed diesel

combustion with high thermal efficiency is so attractive. Recently, the premixed diesel combustion

with high volatility and high auto-ignition resistance fuel like gasoline at the injection timing near

TDC was paid more and more attentions. It has potential to extend the operational load range of

premixed diesel combustion to full load range and the ignition timing can be directly controlled by

varying the injection timing. Furthermore, the possibility of wall wetting and oil dilution would be

decreased with high volatility fuel, resulting in improvement in thermal efficiency of premixed diesel

combustion. However, the effects of fuel properties on premixed diesel combustion with this strategy

are still not so clear and maybe there are the optimum fuel properties including ignitability, volatility,

and compositions for premixed diesel combustion.

The objectives of this study are as follows.

1. To further understand the effects of fuel properties including ignitability, volatility, and

compositions on the stable operational range, combustion characteristics and emissions of

premixed diesel combustion

2. To further understand the reasons and the extent of improvement in thermal efficiency of

premixed diesel combustion with high volatility fuel

3. To obtain the optimized fuel properties for premixed diesel combustion


17

1.4 Layout of dissertation

This dissertation is consisted of seven chapters. The first chapter describes the introduction which

includes general review of the challenge to future internal combustion engines, the principle of

premixed diesel combustion, literature review of past and current trends of premixed diesel

combustion in diesel engines, research objective, and synopsis of each chapter.

Chapter 2 explains experimental apparatus, properties of the tested fuels, procedures, and methods of

data calculation in detail.

Chapter 3 describes the influence of fuel ignitability on premixed diesel combustion with different

ignitability fuels in a supercharged diesel engine with external cooled EGR. Firstly, the base stable

operational range, combustion characteristics and emissions of premixed diesel combustion with an

ordinary diesel fuel were described. Then, the results of four primary reference fuels (PRFs) with

octane numbers from 0 to 60 were mentioned to explain the effect of ignitability on the stable

operational range, combustion characteristics and emissions of premixed diesel combustion. With

increasing octane number, the combustion phasing is retarded and higher intake oxygen concentration

can be used, resulting in higher indicated thermal efficiency due to the higher combustion efficiency

and the larger maximum IMEP of the premixed diesel combustion. Silent, low NOX, and smokeless

operation with a high thermal efficiency is possible up to 1.0 MPa IMEP using PRF with a research

octane number of 40 when the intake oxygen concentration was optimized corresponding to IMEP.

Chapter 4 describes the influence of fuel compositions by investigation with the difference between

PRF and normal heptane and toluene blend fuel (NTF) with the same research octane number as the

PRF, and three research octane numbers of PRF and NTF of 20, 40, and 60. The result shows that, the

operational range of premixed diesel combustion with normal heptane and toluene blend fuels (NTF)

with the same research octane numbers shifts to the higher intake oxygen concentration side,

compared with PRF, showing lower ignition characteristics than PRF.

Chapter 5 describes the influence of fuel volatility on premixed diesel combustion with Normal

heptane, PRF20 and an ordinary diesel fuel (JIS No. 2). Results show that the fuel volatility has little

effect on the stable operational range and combustion characteristics of premixed diesel combustion.

In terms of thermal efficiency, the best indicated thermal efficiency with the ordinary diesel fuel is

lower than with PRF20 while the combustion characteristics with these fuels are similar in a wide

IMEP range. This result is due to the increased quantity of fuel which doesn’t contribute to the

effective combustion, as the high distillation components in diesel fuel adhere to the wall of the

combustion chamber without evaporating.

Chapter 6 describes the influence of fuel volatility on the thermal efficiency of premixed diesel


18

combustion with three types of ordinary diesel fuels with different distillation temperature

distributions (JIS special No.1 diesel fuel, JIS No.1 diesel fuel, and JIS special No.3 diesel fuel) and

PRF20 at different injection timings, coolant temperatures and intake gas temperatures. With the

premixed diesel combustion mode, regardless of injection timings, the indicated thermal efficiencies

with the ordinary diesel fuels were lower than with PRF20 due to the higher unknown losses. With the

conventional diesel combustion mode, the indicated thermal efficiencies with the ordinary diesel fuels

and PRF20 were similar.

Chapter 7 summarizes the results of this research regarding the influence of fuel properties on

operational range, combustion characteristics, emissions and thermal efficiency of premixed diesel

combustion in a diesel engine.


CHAPTER 2. EXPERIMENTAL APPARATUS AND

PROCEDURE


19

2.1 Experimental apparatus

Experiments were conducted on a single cylinder, four-stroke cycle, and direct injection experimental

diesel engine with a supercharger and a common rail fuel injection system. The specifications of the

engine are shown in Table 2-1. In this experiment, supercharging is realized with a compressor driven

by an electric motor and the exhaust manifold pressure was kept constant at the same pressure as the

intake pressure with a throttle valve in the exhaust manifold after one exhaust chamber. For all

experiments in this study, the boost pressure was set at 160 kPa (absolute). The cooled low pressure

loop EGR was realized by diverting a part of the exhaust gas via a DPF and an EGR cooler to the

upstream of a supercharger as illustrated in Figure 2-1. Here, the intake gas temperature was

controlled constant with an intercooler and a heater after the supercharger.

Table 2-1 Test engine specifications

Honda Engine

Type Direct injection, four stroke

Number of cylinders 1

Bore × stroke 85 × 96.9 mm

Displacement 550 cm3

Compression ratio 16.3

Fuel injection system Common rail system

Figure 2-1 Experimental arrangements


20

2.2 Properties of the test fuels

In this experiment, the properties of the tested fuels for the influence of fuel properties on the

operational range and combustion characteristics of premixed diesel combustion are shown in Table

2-2.The effect of fuel ignitability was discussed with four primary reference fuels (PRFs) with octane

numbers from 0 to 60. The effect of the fuel compositions was investigated with the difference

between PRF and normal heptane and toluene blend fuel (NTF) with the same research octane number

as the PRF, and three research octane numbers of PRF and NTF of 20, 40, and 60 were examined. The

effect of fuel volatility was investigated with the difference between normal heptane and an ordinary

diesel fuel (JIS No. 2). In the Table 2-2, the numbers after PRF show the octane numbers. The

numbers after NTF show the fractions of toluene, and the research octane numbers of NTF17, NTF34,

and NTF50 are 20, 40, and 60 respectively. A lubricity improver (R655) was added to PRF and NTF

in a volume fraction of 300 ppm.

Table 2-2 Properties of the tested fuels for the influence of fuel properties on the operational range

and combustion characteristics of premixed diesel combustion in chapter 3, chapter 4, and chapter 5

a 50% distillation temperature

b RON = Toluene content ×1.2

c Cetane number = 60 – RON/2

d PRF: Primary reference fuel for octane number

e NTF: Normal heptane and toluene blend fuel

Table 2-3 Properties of the tested fuels for the effect of fuel volatility on the thermal efficiency of the

premixed diesel combustion in chapter 6

Properties

JIS grade 2 Diesel (DF2)

PRF0d PRF20 PRF40 PRF60 NTF17

e NTF34 NTF50

n-Heptane fraction [%] – 100 80 60 40 83 67 50

Toluene fraction [%] – – – – – 17 34 50

Density [g/cm3] 0.828 0.684 0.685 0.686 0.694 0.718 0.748 0.777

Boiling point [ºC] 175 ~ 354

(286a)

99.4 – – – – – –

Octane number (RONb) – 0 20 40 60 20

b 40

b 60b

Cetane number 54.2 60c 50

c 40

c 30

c 50

c 40

c 30c

Carbon [wt%] 86.2 83.90 83.95 83.99 84.03 85.12 86.34 87.57

Hydrogen [wt%] 13.8 16.10 16.05 16.01 15.97 14.87 13.65 12.42

Stoichiometric ratio 14.5 15.16 15.15 15.14 15.13 14.90 14.62 14.34

LHV [MJ/kg] 43.02 44.56 44.51 44.47 44.43 43.95 43.35 42.74

Fuels


21

Properties SDF1 DF1 SDF3 PRF20

Density [g/cm3] 0.831 0.8342 0.8094 0.685

10% Distillation temperature [ºC] – 217 174

100a 50% Distillation temperature [ºC] – 281 220

90% Distillation temperature [ºC] (≤ 360b) 342

307（≤330

b）

Carbon [wt%] 86.53 86.2 86.24 83.95

Hydrogen [wt%] 13.47 13.8 13.76 16.05

Stoichiometric ratio 15.41 14.66 14.65 15.15

LHV [MJ/kg] 42.96 42.9 43.2 44.51

Octane number (RON) – – – 20

Cetane number 54.2 54.2 49.4 –

a Boiling point b JIS K2204

Table 2-3 shows the properties of the tested fuels for the influence of fuel volatility on the thermal

efficiency of the premixed diesel combustion. It was investigated with the three types of ordinary

diesel fuels: JIS special No.1 diesel fuel (SDF1), JIS No.1 diesel fuel (DF1), and JIS special No.3

diesel fuel (SDF3) with different distillation temperature distributions, and a primary reference fuel

with the octane number of 20 (PRF20). In Japanese Industrial Standard (JIS K2204) SDF1 is the

highest distillation temperature grade and SDF3 is the lowest distillation temperature grade among

diesel fuels. The influence of fuel volatility on the thermal efficiency of conventional diesel

combustion was also examined with SDF1 and PRF20.

2.3 Fuel injection systems

2.3.1 Nozzle specifications

Table 2-4 shows nozzle specifications for this experiment.

Table 2-4 Nozzle specifications

Type Sac nozzle

Spray angle 156º

Number of nozzle hole 7

Hole diameter (mm) 0.125

2.3.2 Common rail systems

A common rail fuel injection system illustrated in Figure 2-2 is used in this experiment.

Fuels


22

The injection pressure with a common rail system is independent on engine speed and fuel supply

quantities, which is different from the conventional system. The main elements of the common rail

systems are an adjustable high pressure fuel pump, a common rail, an injector, and an accurate control

system of ECU. As the fuel pump delivers variable fuel pressures to common rail and accurately

controls the rail pressure through the pressure sensor, the pressure in common rail is stable. One

pressure limiter is attached to avoid excess pressure in the common rail. This adjustable fuel pressure

in the common rail is then delivered to injectors which are equipped with solenoids to control the

injection timings and quantities. Here, the injection timing and quantity are controlled with ECU.

While the maximum injection pressure with this system is 180 MPa, the injection pressure is mainly

set at 160 MPa in this experiment.

Figure 2-2 The common rail fuel injection system [52]

2.3.3 Injection strategy

In the investigation of the effect of fuel properties on the operational range and combustion

characteristic of premixed diesel combustion in chapter 3, chapter 4, chapter 5, the fuel injection

timing was always set at 25ºCA BTDC. With this fuel injection timing, the fuel spray impinges on the

tip of the piston bowl. In the investigation of fuel volatility on the thermal efficiency of the premixed

diesel combustion, three fuel injection timings of 40ºCA BTDC, 25ºCA BTDC, and 10ºCA BTDC

were set for the premixed diesel combustion mode, while a pilot fuel injection timing of 8ºCA BTDC,

and a main injection timing of 3ºCA ATDC were used for the conventional diesel combustion mode.

Figure 2-3 shows the positions of the piston and fuel sprays at the different fuel injection timings for

the premixed diesel combustion mode. With the injection timing of 40ºCA BTDC, fuel sprays do not

enter the piston bowl and directly impinge on the cylinder liner wall while with the injection timings

of 25ºCA BTDC and 10ºCA BTDC, almost all of the fuel sprays enter the combustion chamber.


23

Figure 2-3 Positions of the piston and fuel spray at different fuel injection timings under the premixed

diesel combustion mode

2.4 Operating conditions

2.4.1 The operating conditions for fuel properties on operational range and combustion

characteristics of premixed diesel combustion

Table 2-5 The operating conditions for fuel properties on operational range and combustion

characteristics of premixed diesel combustion in chapter 3, chapter 4, and chapter 5

Engine speed 2000 rpm

Fuel injection pressure 160 MPa

Fuel injection timing 25ºCA BTDC

Boost pressure 160 kPa (abs)

Intake gas temperature 25ºC

Coolant temperature 80ºC

lubricant temperature 80ºC

The operating conditions in chapter 3, chapter 4, and chapter 5 are shown as Table 2-5. The engine

speed was always set at 2000 rpm, the fuel injection pressure at 160 MPa. The fuel injection timing

was always set at 25ºCA BTDC. The boost pressure was set at 160 kPa (absolute) with a supercharger


24

driven by an electric motor, and an electric heater and an intercooler were installed after the

supercharger to maintain the intake gas temperature at 25ºC. The coolant and lubricant temperatures

were both set at 80ºC.

2.4.2 The operating conditions for fuel distillation temperature on thermal efficiency of

premixed diesel combustion in chapter 6

The operating conditions for fuel distillation temperature on thermal efficiency of premixed diesel

combustion in chapter 6 are shown in Table 2-6. An intake oxygen concentration of 11%, an

equivalence ratio of 0.5 and three fuel injection timings of 40ºCA BTDC, 25ºCA BTDC, and 10ºCA

BTDC were set for the premixed diesel combustion mode, while an intake oxygen concentration of

21%, an IMEP of 0.7 MPa, a pilot fuel injection timing of 8ºCA BTDC, and a main injection timing of

3ºCA ATDC were used for the conventional diesel combustion mode. With the injection timing of

40ºCA BTDC, fuel sprays do not enter the combustion chamber and directly impinge on the cylinder

liner wall while with the injection timings of 25ºCA BTDC and 10ºCA BTDC, almost all of the fuel

sprays enter the combustion chamber.

The coolant temperature was maintained at 80ºC and also investigated at 40ºC for the premixed diesel

combustion mode. The intake gas temperature was usually maintained at 40ºC and temperatures of

25ºC and 70ºC were also investigated for the premixed diesel combustion mode.


25

Table 2-6 The operating conditions for fuel distillation temperature on thermal efficiency of premixed

diesel combustion in chapter 6

Engine speed 2000 rpm

Fuel injection pressure 160 MPa

Boost pressure (abs.) 160 kPa

Premixed diesel combustion mode

Equivalence ratio 0.5

Intake oxygen concentration 11%

Fuel injection timing 40º, 25º,10ºCA BTDC

Intake gas temperature 25º, 40º, 70ºC 40ºC

Coolant temperature 80ºC 40ºC

Lubricant oil temperature 80ºC 70ºC

Conventional diesel combustion mode

IMEP 0.7 MPa

Intake oxygen concentration 21%

Fuel injection timing Pilot 8ºCA BTDC

Main 3ºCA ATDC

Intake gas temperature 40ºC

Coolant temperature 80ºC

Lubricant oil temperature 80ºC

2.5 Measurement and analysis of indicator diagram

2.5.1 Acquisition of in-cylinder pressure

The cylinder pressure was measured with a piezo type of pressure transducer (KISTLER 6052C)

equipped next to the injector in the cylinder head. The transducer has a range between 0 and 25 MPa,

and designed for durable from engine knocking. The electric charge from the pressure transducer then

amplified and converted into voltage with a KISTLER 5011B type charge amplifier. The voltage then

can be monitor with a digital oscilloscope (Tektronix TDS2014) and is simultaneously converted to

digital signal through A/D converter (Interface CBI-320412) connected to a computer.

Figure 2-4 shows the in-cylinder pressure acquisition system in the experiment. In addition to monitor

in-cylinder pressure, a digital oscilloscope also was used for monitoring the injection duration and

timing signal from the injector solenoid valve, and the crank angle signal via phototransistor attached


26

at the disc equipped between the engine and the dynamometer. The disc itself has 1 mm slit per 20

degrees and phototransistor (NEC PD32) detected the light striking on it then converted to per 20

degrees of crank angle signal. A rotary encoder (TAMAGAWA SEIKI TS5011N) was attached

directly on crank shaft for the pressure data sampling. The trigger signal was set at 45ºCA BTDC

(Before Top Dead Center) to start the sampling and the clock signal per 0.2 degree crank angle was

sent to computer. This digitized signal has 1800 recorded in-cylinder pressure every one engine

rotation. The 200 cycles were averaged and used for calculation of the rate of heat release.

Figure 2-4 In-cylinder pressure acquisition system

2.5.2 Heat release calculation methodology

The in-cylinder pressure data was used for calculating the rate of heat release and the degree of

constant volume heat release to analyze its combustion behavior. Considering the changes in

temperature and composition of working gas in cylinder, the working gas was treated as semi perfect.

From the first law of thermodynamics,

dQ dU dVP

d d d = ( )V

d dVnC T P

d d [2-1]

And where

Q : Heat release in the cylinder [J]

U : Internal energy of working gas [J]

V : Cylinder volume [m3]

CV : Specific heat at constant volume of working gas [J/mol⋅K]

n : Mol number of working gas [mol]

T : Average temperature of working gas [K]


27

P : In-cylinder pressure [Pa]

: Crank angle [CA]

Here,

1

V

RC

[2-2]

R : Universal gas constant [J/mol⋅K]

: Ratio of specific heat of working gas

Then from PV=nRT, dQ

d can be expressed as follows:

2

1

1 ( 1)

dQ dP dV PV dV P

d d d d

[2-3]

The average specific heat of in-cylinder gas can be established by the equation of Kühl or JANAF. In

this calculation, the equation of Tanishita is used to obtain . First of all, based on Tanishita’s heat

specific equation, which is:

2100

C100 100

P

T TC A B D

T

[2-4]

Table 2-7 Constant of Tanishita’s specific heat equation [53]

Gas Temperature

Range [K] A B

H2O 300-900 26.523378 1.3640594

900-3000 28.652784 1.6501498

O2 300-900 17.579533 2.7102663

900-3000 38.806193 -0.0108814

CO2 300-900 34.242162 3.4580037

900-3000 63.617588 0.1972192

N2 300-900 24.084567 0.7549177

900-3000 35.00709 0.242445

Gas Temperature

Range [K] C D

H2O 300-900 0 8.6415552

900-3000 -0.0254607 -15.531939


28

O2 300-900 -0.1128853 14.081338

900-3000 0.0031259 -40.920401

CO2 300-900 -0.1254832 -18.626528

900-3000 -0.0035441 -107.5861

N2 300-900 0.0043693 8.3191719

900-3000 -0.0042136 -40.361623

Gas constants A, B, C, and D for various gas components are shown in Table 2-7. The specific heat at

constant pressure of Cp for every component is determined, and can be calculated as follows:

p

p

C

C R

[2-5]

2.5.3 The degree of constant volume heat release

The degree of constant volume heat release can be defined as the ratio of theoretical work obtained

from heat at each of crank angle and theoretical work of Otto cycle. According to energy balance, the

relation among the thermal efficiency related parameters can be expressed as the following equation:

i = glh · u · (1 – w) · th [2-6]

Where, th : The theoretical thermal efficiency of the Otto cycle

i : The indicated thermal efficiency calculated with the fuel consumption per cycle and the

in-cylinder pressure

glh : The degree of constant volume heat release obtained from the rate of heat release

u : The combustion efficiency calculated from the carbon monoxide (CO) and total

hydrocarbon (THC) concentrations in the exhaust gas

Figure 2-5 shows the indicator diagram and the segmental Otto cycle which bounded by many small

constant volume line. The thermal efficiency of this segmental Otto cycle can be represented by:

1

11th

[2-7]

Here,1V

V

with V1 is the cylinder volume at the bottom dead center and Vθ is the cylinder volume

at heat release.

is the specific heat ratio at the heat release. Hence, by defined dQ

d as the rate of

heat release, the total work of the segmental Otto cycles can be represented by,


29

Figure 2-5 Segmental Otto cycle

b

e tha

dQW d

d

[2-8]

Where, a : Crank angle at the start of heat release

b : Crank angle at the end of heat release

Then, the theoretical thermal efficiency of Otto cycle can be calculated from compression ratio as

follow.

11 (1 )th

[2-9]

Here, ε is the geometrical compression ratio of test engine and is the average specific heat ratio

from the start of injection to the end of heat release.

The degree of constant volume heat release can be calculated as follows.

(1 )b

glh th tha

dQQ d

d

[2-10]

Here, b

a

dQQ d

d

is the total heat release.

2.6 Analysis of exhaust gas emissions

The exhaust gas were sampled at 2200 mm downstream from the exhaust port to measure the THC,

CO, CO2, NOX, and smoke emissions. The CO and CO2 were analyzed with an automotive exhaust gas

analyzer (MEXA-9100DEGR: HORIBA) using a NDIR (non-dispersive infrared absorption) sensor.

The measurement range for CO and CO2 are 3% and 10%, and 6% and 16%, respectively. The THC

and NOX were analyzed with a HFID (Heated flame ionization detector) and a CLD (chemical

V

P

V2

V

dQ

V1

V

P

V2

V

dQ

V1


30

luminescence detector), respectively. This analyzer has a range between 50 ppm and 25000 ppm for

THC, and 20 ppm and 10000 ppm for NOX.

Smoke density was measured with a Bosch-type smoke meter (MODEL GSM-2: SOKKEN) from the

exhaust gas sampled at about 2200 mm downstream from the exhaust port. The measurement range of

the smoke meter is from -3.5% until 100% of smoke density which expressed in Bosch smoke unit. In

this experiment, zero point was always set at -3.5% before sampling measurements.

2.7 Other measurements

An eddy current dynamometer (SEIDENSHA SB-22769) as shown in Figure 2-6 was used to measure

the torque and to absorb power from test engine. It used external supplied water for coolant.

Figure 2-6 Dynamometer

One 4-radial piston type, high accuracy flow meter (MODEL 213/294: TOYO CONTROLS CO. LTD)

as shown in Figure 2-7 was used for measuring fuel consumption. Fuel consumption is determined by

the time for a volume of 45 mm3. Returned fuel from the common rail and the injector were cooled by

external water to keep the inlet fuel temperature constant.

Figure 2-7 Fuel flowmeter

The engine speed was measured by a speed sensor installed in the engine and the signal is collected by

INCA (Integrated Calibration and Acquisition systems). Then the engine speed can be changed by

speed instrument. The engine speed was set at 2000 rpm for all experiments.


31

Intake air quantity was measured with pressure difference at an orifice located before a surge tank as

shown in Figure 2-8. Here, the orifice has a diameter of 37.8 mm and a discharge coefficient of 0.82.

The pressure difference was measured by a digital manometer (PE-33-D) as shown in Figure 2-9.

Figure 2-8 Surge tank

Figure 2-9 Digital manometer

The oxygen concentrations were measured with a portable oxygen tester (POT-101: SHIMAZU) using

thermo-magnetic sensor with a range of 10% and 25% as shown in Figure 2-10. In this oxygen

analyzer, pure nitrogen is used for zero setting and surrounding air for span check. Intake oxygen

concentration was measured at 90 mm upstream from the end of the intake manifold and exhaust

oxygen concentration was measured at 2200 mm downstream from the front of the exhaust manifold.

Figure 2-10 Portable oxygen tester (POT-101: SHIMAZU)

The coolant temperature was measured with a K type thermocouple (CHINO, KT27A). This

temperature sensor was placed between the coolant tank outlet and the inlet pump pipe. OMRON

E5CN type temperature controller and a heater were employed to keep the inlet coolant temperature

constant.

Intake gas temperature was controlled by changing the flow rate of the water through the intercooler

installed at the downstream of supercharger and the target temperature of the heater. And Intake gas


32

temperature was measured with a K type thermocouple (TEMPERATURE CONTROLLER FHP-201:

OMRON) at 50 mm upstream from the end of the intake manifold. The exhaust gas temperature was

measured with a temperature controller (E5CN type: OMRON) at 50 mm downstream from the front

of the exhaust manifold.

Intake gas pressure was measured with a MAP/TA (Manifold Air Pressure/Air Temperature) sensor

installed at 50 mm upstream from the end of the intake manifold. The exhaust pressure was measured

with a pressure gauge. The exhaust pressure was measured at 50 mm downstream the front of the

exhaust manifold.


CHAPTER 3. INFLUENCE OF FUEL IGNITABILITY

ON OPERATIONAL RANGE AND COMBUSTION

CHARACTERISTICS OF PREMIXED DIESEL

COMBUSTION


33

3.1 Chapter outline

In this experiment the influence of fuel ignitability on premixed diesel combustion was examined.

Firstly, the results with JIS No.2 diesel fuel was examined to show the fundamental characteristics of

premixed diesel combustion in this experimental condition. Four primary reference fuels for octane

numbers from 0 to 60 were examined to investigate the influence of fuel ignitability on the operational

range, combustion characteristics, emissions and thermal efficiency of premixed diesel combustion.

Moderately advanced injection timing at25oCA BTDC was set as well as high injection pressure at

160MPa to promote the premixing process before the combustion starts. High EGR rates were also

applied to suppress the pressure rise rate and control NOX emissions in low level.

3.2 The stable operational range and combustion characteristics of premixed diesel combustion

with diesel fuel

Figure 3-1 shows the range of intake oxygen concentrations, O2in, for stable premixed diesel operation,

and the excess air ratio with ordinary diesel fuel (JIS No. 2) for various IMEP (indicated mean

effective pressure). The operation becomes impossible due to knocking (the rate of pressure rise

higher than 1.0 MPa/oCA) when the oxygen concentration exceeds the solid line, and also when the

oxygen concentration decreases and reaches the blue dashed line, here due to unstable combustion (the

coefficient of variance (COV) in IMEP exceeds 5%) below 0.65 MPa IMEP and at the green dashed

line due to deterioration of the combustion efficiency (the indicated thermal efficiency below 40%)

above 0.65 MPa IMEP. Further, with increases in IMEP, the stable operational range of intake oxygen

concentrations between the knocking and stable limits becomes narrower.

The square symbols and the solid lines in Figure 3-2 show the optimum intake oxygen concentrations,

O2in, and excess air ratios where the best indicated thermal efficiency is obtained in the stable

premixed diesel operational range shown in Figure 3-1. The optimum intake oxygen concentrations for

the indicated thermal efficiencies are always obtained relatively near the knocking limit line for the

various IMEP.


34

Figure 3-1 Practical premixed diesel operational range of intake oxygen concentrations, O2in, and

excess air ratios with JIS No.2 diesel fuel

Figure 3-2 Optimum intake oxygen concentrations, O2in, and excess air ratios for the best indicated

thermal efficiencies with JIS No.2 diesel fuel


35

Figure 3-3 shows thermal efficiency related parameters including the indicated thermal efficiency, i,

the degree of constant volume heat release, glh, the combustion efficiency, u, and the unknown loss,

w with ordinary diesel fuel at the optimum thermal efficiency conditions shown in Figure 3-2.

Here, the unknown loss, w is calculated by the following equation based on the heat balance.

i = glh · u · (1 – w) · th

th：The theoretical thermal efficiency of the Otto cycle

i ：The indicated thermal efficiency calculated by the fuel consumption per cycle and the

in-cylinder pressure

glh：The degree of constant volume heat release obtained from the rate of heat release

u：The combustion efficiency calculated from the CO and THC concentrations in the exhaust gas

The unknown loss, w includes the following:

1. Cooling loss (main component)

2. Unburned fuel mixed into the lubricant oil

3. Oxidized fuel after the apparent heat release in the cylinder during the expansion and exhaust

strokes as well as in the exhaust manifold before the exhaust gas sampling

As shown in Figure 3-3, high indicated thermal efficiencies can be obtained at all IMEP conditions

except above 0.8 MPa IMEP, where the combustion efficiency deteriorates markedly. Here a much

higher degree of constant volume heat release and smaller unknown loss (mainly cooling loss) than

ordinary diesel combustion is realized; however the combustion efficiency is very low. This is a

general characteristic of premixed diesel combustion.


36

Figure 3-3 Thermal efficiency related parameters with diesel fuel at the optimum intake oxygen

concentrations

Figure 3-4 shows the exhaust gas emissions and the maximum rates of pressure rise at the optimum

intake oxygen concentrations for the thermal efficiencies shown in Figure 3-2. The NOX emissions are

everywhere below 20 ppm and smokeless operation is realized under all these conditions. The levels

of CO and THC emissions are relatively low below 0.8 MPa IMEP, but increase significantly above

0.8 MPa IMEP when the excess air ratio is close to unity and the operation becomes impossible, even

though the rate of pressure rise and COV are sufficiently low.


37

Figure 3-4 Exhaust gas emissions and the maximum rates of pressure rise, dp/dmax with diesel fuel at

the optimum intake oxygen concentrations

3.3 Effect of fuel ignitability on stable operational range of premixed diesel combustion

Figure 3-5 shows the stable premixed diesel operational range of intake oxygen concentrations, O2in

with PRF0 (n-heptane) and PRF60 for various IMEP. With increases in the octane number from 0 to

60, both the knocking and stable limits shift to higher intake oxygen concentrations, and the maximum

IMEP of the premixed diesel combustion increases from 0.83 MPa to 1.0 MPa. However, the stable

operational range of intake oxygen concentrations between the knocking and stable limits is narrower


38

with PRF60 than with PRF0 and decreases with increases in IMEP, resulting in a reduction of

combustion control robustness.


and PRF60

Figure 3-6 shows the in-cylinder pressure and the rate of heat release (ROHR) at 0.4 MPa IMEP with

PRF0 at an oxygen concentration of 16.2% and PRF60 at an oxygen concentration of 16.5%. Even

though the oxygen concentration of PRF60 is higher than of PRF0, with PRF60 the ignition timing

and combustion phase are retarded and the heat release is less abrupt (milder) than with PRF0. These

ignition and combustion characteristics of PRF60 can increase the maximum IMEP with premixed

diesel operation as shown in Figure 3-5.

Figure 3-7 shows the in-cylinder pressure and the rate of heat release (ROHR) at 0.6 MPa IMEP with

PRF0 at an oxygen concentration of 11.4% and PRF60 at an oxygen concentration of 17%. These

combustion characteristics are very similar, showing that premixed diesel combustion obtained at

lower oxygen concentrations with higher ignitability fuels can also be realized at higher intake oxygen

concentrations with lower ignitability fuels.


39


PRF60 at O2in = 16.5% (IMEP: 0.4 MPa)


PRF60 at O2in = 17% (IMEP: 0.6 MPa)


40

3.4 Effect of fuel ignitability on combustion characteristics and emissions of premixed diesel

combustion

Figure 3-8 shows the optimum intake oxygen concentrations for the thermal efficiency, the optimum

excess air ratios, and the stable premixed diesel operational range with PRF0 (normal heptane) and

PRF40 plotted versus IMEP. With the higher octane number (40 versus 0), the optimum intake oxygen

concentration and excess air ratio for the best indicated thermal efficiency as well as the stable

operational range becomes higher. At both octane numbers, the optimum intake oxygen concentration

is obtained near the knocking limit.

Figure 3-8 The optimum intake oxygen concentrations, O2in and the excess air ratios for the best

indicated thermal efficiencies with PRF0 and PRF40

Figure 3-9 shows thermal efficiency related parameters including the indicated thermal efficiency, i,

the degree of constant volume heat release, glh, the combustion efficiency, u, and the unknown loss,

w with PRF0 and PRF40 at the optimum intake oxygen concentrations shown in Figure 3-8. While

there is little difference in the indicated thermal efficiencies at low loads, the indicated thermal

efficiency with PRF40 is better than with PRF0 at high loads due to the higher combustion efficiency,


41

while the degrees of constant volume heat release are similar and very high. Particularly, the very high

indicated thermal efficiency of 48% is obtained with PRF40 at 0.8 MPa IMEP due to the relatively

high combustion efficiency, smaller unknown loss, and the high degree of constant volume heat

release.

Figure 3-9 Thermal efficiency related parameters with PRF0 and PRF40 at the optimum intake

oxygen concentrations

Figure 3-10 shows the combustion related parameters, 10% heat release crank angle, CA10, 50% heat

release crank angle, CA50, the premixing duration, pre, and the combustion duration, comb (crank

angle between CA10 and CA90); with PRF0 and PRF40 at the optimum intake oxygen concentrations


42

for the thermal efficiencies shown in Figure 3-8. Here, CA10 corresponds to the ignition timing and

the premixing duration, pre is defined as the crank angle from the end of fuel injection to CA10. Due

to the lower ignitability of PRF40, the ignition timings with PRF40 are more retarded than with PRF0

at high IMEP conditions, resulting in longer premixing durations regardless of the higher optimum

intake oxygen concentrations, while there is little difference in ignition timings with PRF0 and PRF40

at low IMEP condition. The 50% heat release crank angles, CA50 are near the top dead center and the

combustion durations are very short in all cases, resulting in the very high degree of constant volume

heat release shown in Figure 3-9.


43

Figure 3-10 Combustion related parameters with PRF0 and PRF40 at the optimum intake oxygen

concentrations

Figure 3-11 shows the exhaust gas emissions and the maximum rates of pressure rise with PRF0 and

PRF40 at the optimum intake oxygen concentrations shown in Figure 3-8. Smokeless and very low

NOX combustion is realized at the optimum intake oxygen concentrations over all of the IMEP range

here. However, the CO and THC emissions for both fuels are still very high especially at high IMEP

due to the lower oxygen concentrations and excess air ratios, resulting in the deterioration of the

combustion efficiencies. Especially, the CO and THC emissions from PRF0 is higher than from

PRF40 as the intake oxygen concentrations have to be decreased more to obtain the best thermal


44

efficiencies. There is little difference in the maximum rate of pressure rise of PRF0 and PRF40 where

the best thermal efficiencies are obtained.

Figure 3-11 Exhaust gas emissions and the maximum rates of pressure rise with PRF0 and PRF40 at

the optimum intake oxygen concentrations

3.5 Chapter summary

In this chapter, four primary reference fuels for octane numbers from 0 to 60 were used to investigate

the influence of fuel ignitability on the operational range, combustion characteristics, emissions and

thermal efficiency of premixed diesel combustion.


45

The results may be summarized as follows:

1. The IMEP of the premixed diesel combustion is limited by knocking with high intake oxygen

concentrations and by unstable combustion or significant increases in CO and THC emissions with

low intake oxygen concentrations regardless of the fuels examined in this experiment.

2. With increasing octane number of the primary reference fuels (PRF), the combustion phasing is

retarded, and higher intake oxygen concentrations can be used within the tolerance limits of the

rapid combustion, resulting in the higher maximum IMEP of the premixed diesel combustion and

higher indicated thermal efficiencies due to the higher combustion efficiency. However, the stable

operational range of intake oxygen concentrations between the knocking and stable limits is

narrower with high octane number fuels and decreases with increases in IMEP, resulting in a

reduction of combustion control robustness.

3. Regardless of the fuels tested in this research, the optimum intake oxygen concentration for the

best thermal efficiency is obtained near the knocking limit over a wide IMEP ranges.

4. The premixed diesel combustion which is obtained at lower oxygen concentrations with higher

ignitability fuels can be realized at higher intake oxygen concentrations with lower ignitability

fuels.

5. Silent, low NOX, and smokeless operation with a high thermal efficiency was possible up to 1.0

MPa IMEP with PRF with a research octane number of 40 when the intake oxygen concentration

was optimized corresponding to IMEP and the boost pressure was set at 160 kPa (absolute).

6. In all of the IMEP range, even at the optimum intake oxygen concentrations for the thermal

efficiency with the premixed diesel combustion mode, the levels of CO and THC emissions are

still much higher than that with the ordinary diesel combustion mode, regardless of the fuels

examined in these experiments.


CHAPTER 4. INFLUENCE OF FUEL

COMPOSITIONS ON OPERATIONAL RANGE AND

COMBUSTION CHARACTERISTICS OF

PREMIXED DIESEL COMBUSTION


46

4.1 Chapter outline

In this experiment the influence of the fuel compositions on the operational range, combustion

characteristics, emissions and thermal efficiency of premixed diesel combustion was investigated with

the difference between PRF and normal heptane and toluene blend fuel (NTF) with the same research

octane number as the PRF, and three research octane numbers of PRF and NTF of 20, 40, and 60 were

examined. The numbers after PRF show the octane numbers and the numbers after NTF show the

fractions of toluene, and the research octane numbers of NTF17, NTF34, and NTF50 are 20, 40, and

60 respectively. Due to the important effect of the fuel chemical characteristics on low temperature

oxidation in premixed diesel combustion mode, so this experiment is necessary. Moderately advanced

injection timing at 25oCA BTDC was also set as well as high injection pressure at 160MPa to promote

the premixing process before the combustion starts. High EGR rate was also applied to suppress the

pressure rise rate and control NOX emissions in low level.

4.2 Effect of fuel compositions on stable operational range of premixed diesel combustion

Figure 4-1, Figure 4-2 and Figure 4-3 show the stable premixed diesel operational range of intake

oxygen concentrations, O2in with PRF20 and NTF17, PRF40 and NTF34 and PRF60 and NTF50. Here,

PRF20 and NTF17, PRF40 and NTF34 and PRF60 and NTF50 have the same research octane

numbers of 20, 40 and 60. As these three figures show, the smooth operational range with NTF is

similar to the PRF with the same research octane number. However, the knocking and stable limits

with NTF shift to higher intake oxygen concentrations than with PRF with the same research octane

number. These results show that NTF has lower ignitability characteristics than PRF with the same

research octane number.


47


and NTF17


and NTF34


48


and NTF50

Figure 4-4 shows the maximum IMEP and the intake oxygen concentration when the maximum IMEP

is obtained with PRF and NTF. The maximum IMEP with NTF is slightly higher than with PRF with

the same research octane numbers due to the longer ignition delays.

Figure 4-4 Effect of octane number on the maximum IMEP and the intake oxygen concentration at the

maximum IMEP for PRF and NTF


49

Figure 4-5 shows the in-cylinder pressure and the rate of heat release (ROHR) with PRF40 and NTF34

under the same intake oxygen concentrations of the 14.4% at a 0.6 MPa IMEP condition. The ROHR

during low temperature oxidation is enlarged at the right side of the plots with two abscissas, for crank

angle and in-cylinder gas temperature. With NTF34 the onset timing of the low temperature oxidation

is retarded, the temperature at low temperature oxidation onset rises, and the rate and quantity of the

heat release with low temperature oxidation decrease below those of PRF40 with the same research

octane numbers, resulting in a more retarded and slower main combustion. This is due to the inhibitor

effect on low temperature oxidation of toluene [54] and also due to the higher ignition temperature of

toluene itself [55]. Similar results were obtained at both 0.4 MPa and 0.8 MPa IMEP conditions.

Figure 4-5 In-cylinder pressure and the rate of heat release (ROHR) with PRF40 and NTF34 under the

same intake oxygen concentrations of 14.4% (IMEP: 0.6 MPa)

4.3 Effect of fuel compositions on combustion characteristics and emissions on premixed diesel

combustion

The following will compare the parameters related to thermal efficiency and exhaust gas emissions for

a primary reference fuel (PRF) and a normal-heptane toluene blend fuel (NTF) under optimum thermal

efficiency conditions in stable premixed diesel combustion. The test fuels are PRF40 and NTF34,

which both have the research octane number of 40. Figure 4-6 shows the oxygen concentration and

excess air ratio when the indicated thermal efficiency reaches the maximum in the stable premixed


50

diesel operational range. The optimum indicated thermal efficiencies are obtained relatively near the

knocking limit line for both fuels, and the intake oxygen concentrations and excess air ratios for the

optimum efficiencies with NTF34 are higher than with PRF40 over a wide IMEP range.

Figure 4-6 The optimum intake oxygen concentration, O2in and the excess air ratio for the best

indicated thermal efficiency with PRF40 and NTF34

Figure 4-7 shows the thermal efficiency related parameters including the indicated thermal efficiency,

i, the degree of constant volume heat release, glh, the combustion efficiency, u, and the cooling loss,

w with PRF40 and NTF34 at the optimum thermal efficiency conditions shown in Figure 4-6. The

indicated thermal efficiencies with PRF40 are slightly higher than with NTF34 except at high IMEPs

in spite of the lower combustion efficiencies. This is due to lower cooling losses with lower intake

oxygen concentrations. The degrees of constant volume heat release are very high between 97.5% and

99.5% for all cases, and very high indicated thermal efficiencies of 48% at 0.8 MPa IMEP are

obtained for both fuels. The combustion efficiencies are much lower than with ordinary diesel

combustion and decrease to below 90% at 0.4 MPa IMEP for both fuels. Thus, there is a possibility to

improve the thermal efficiency of premixed diesel combustion further with increases in combustion

efficiency.


51

Figure 4-7 Thermal efficiency related parameters with PRF40 and NTF34 at the optimum intake

oxygen concentrations

Figure 4-8 shows the combustion related parameters including the 10% heat release crank angle,

CA10, 50% heat release crank angle, CA50, the premixing duration, pre, and the combustion duration,

comb (duration between CA10 and CA90) with PRF40 and NTF34 at the optimum thermal efficiency

conditions shown in Figure 4-6. Here, CA10 corresponds to the ignition timing and the premixing

duration, pre is defined as the crank angle from the end of fuel injection to CA10. While NTF34 has

poorer ignitability as mentioned previously, there is little difference in CA10 between PRF40 and

NTF34 as the intake oxygen concentrations of NTF are higher, to enable obtaining the optimum


52

thermal efficiency. The 50% heat release crank angles, CA50 are near the top dead center and the

combustion durations are very short for all cases, resulting in the very high degree of constant volume

heat release shown in Figure 4-7.

Figure 4-8 Combustion related parameters with PRF40 and NTF34 at the optimum intake oxygen

concentrations

Figure 4-9 shows the exhaust gas emissions with PRF40 and NTF34 at the optimum thermal

efficiency conditions shown in Figure 4-6. Smokeless and very low NOX combustion is realized in a

wide IMEP range up to 1.0 MPa IMEP. The CO and THC emissions for both fuels are very high, and

both kinds of emissions from PRF40 are slightly higher than those from NTF34 due to the lower


53

optimum intake oxygen concentrations. There is no significant difference for the maximum rate of

pressure rise of PRF40 and NTF34 at the values where the best thermal efficiencies are obtained.

Figure 4-9 Exhaust gas emissions and the maximum rates of pressure rise, dp/dmax with PRF40 and

NTF34 at the optimum intake oxygen concentrations

4.4 Chapter summary

In this Chapter, the operational ranges, combustion characteristics, and emissions of premixed diesel

combustion with normal heptane and toluene blend fuels (NTF) and the primary reference fuels (PRF)


54

with similar research octane numbers were compared to show the effect of fuel composition on

premixed diesel combustion, The results may be summarized as follows:

1. The optimum intake oxygen concentrations for the thermal efficiency and for the operational

range with normal heptane and toluene blend fuels (NTF) are higher than with the primary

reference fuels (PRF) with similar research octane numbers, showing the poorer ignition

characteristics than PRF.

2. The onset timings of the low temperature oxidation with normal heptane and toluene blend fuels

are retarded and the rate and quantity of the heat release with low temperature oxidation decreases

to levels lower than PRF with the same octane numbers, resulting in a more retarded and slower

main combustion. This is due to the inhibitor effect on the low temperature oxidation of toluene

and also due to the higher ignition temperature of toluene itself.


MPa IMEP both with PRF and NTF with a research octane number of 40 when the intake oxygen

concentration was optimized corresponding to IMEP and the boost pressure was set at 160 kPa

(absolute).


CHAPTER 5. INFLUENCE OF FUEL VOLATILITY

ON OPERATIONAL RANGE AND COMBUSTION

CHARACTERISTICS OF PREMIXED DIESEL

COMBUSTION


55

5.1 Chapter outline

In this experiment, the influence of fuel volatility on premixed diesel combustion was investigated

with the difference between normal heptane, PRF20 and JIS No. 2 diesel fuel. Due to the important

effect of fuel volatility on premixing process in premixed diesel combustion mode, so this experiment

is necessary. Moderately advanced injection timing at25oCA BTDC was also set as well as high

injection pressure at 160 MPa to promote the premixing process before the combustion starts. High

EGR rate was also applied to suppress the pressure rise rate and control NOX emissions in low level.

5.2 Effect of fuel volatility on stable operational range of premixed diesel combustion

Figure 5-1 shows the stable premixed diesel operational range of intake oxygen concentrations, O2in

with ordinary diesel fuel and normal heptane for various IMEP (indicated mean effective pressure).

While the ignitability of the diesel fuel and normal heptane are similar, the volatilities are significantly

different as the 50% distillation temperature of the diesel fuel is 286ºC and the boiling point of normal

heptane is 99ºC. Operation becomes impossible when the oxygen concentrations exceed the solid lines

due to knocking, and also when the oxygen concentration decrease and reach the dashed lines due to

unstable combustion below 0.65 MPa IMEP or deterioration of the combustion efficiency above 0.65

MPa IMEP. There is little difference in the operational range with diesel fuel and normal heptane with

the similar ignitibilities but different volatilities, showing that fuel volatility has little effect on the

operational range of premixed diesel combustion.

Figure 5-1 Stable premixed diesel operational range of intake oxygen concentrations, O2in

with

ordinary diesel fuel (JIS No.2 diesel fuel) and normal heptane for various IMEP


56

5.3 Effect of fuel volatility on combustion characteristics and emissions on premixed diesel

combustion

Figure 5-2 shows the in-cylinder pressure and the rate of heat release (ROHR) with diesel fuel and

normal heptane at 0.6 MPa IMEP and an oxygen concentration of 11.4%. The combustion

characteristics with both fuels are very similar, showing that fuel volatility has little effect on

premixed diesel combustion when the output of the engine is the same with different fuel volatility.

Figure 5-2 In-cylinder pressure and the rate of heat release (ROHR) with diesel fuel and normal

heptane at O2in

= 11.4% (IMEP: 0.6 MPa)

Figure 5-3 shows the CO, THC, and NOX emissions with diesel fuel and normal heptane for various

IMEP at the knocking limits. The THC with diesel fuel is slightly higher than that with normal heptane

in a wide IMEP range, but there is little difference in the exhaust gas emissions of the two fuels. The

levels of CO and THC emissions are relatively low below 0.8 MPa IMEP, but increase significantly

above 0.8 MPa IMEP where the operation becomes impossible even though the rate of pressure rise

and COV are low. The NOX emissions are everywhere below 30 ppm and smokeless operation is

realized for all these conditions.


57

Figure 5-3 Exhaust gas emissions with diesel fuel and normal heptane for various IMEP at the

knocking limits

Figure 5-4 shows the optimum intake oxygen concentrations for the thermal efficiencies with PRF20

and JIS No.2 diesel fuel versus IMEP. The similar operational load range and optimum intake oxygen

concentrations for the thermal efficiencies show that the ignitability of PRF20 and JIS No.2 diesel fuel

are similar.

Figure 5-4 The optimum intake oxygen concentration, O2in

, for the best thermal efficiency with JIS

No.2 diesel fuel and PRF20


58

Figure 5-5 shows the thermal efficiency related parameters with PRF20 and JIS No.2 diesel fuel

versus IMEP at the optimum intake oxygen concentration condition. While the ignitability of PRF20

and JIS No.2 diesel fuel are similar, the volatilities are significantly different as the boiling point of

PRF20 is about 99ºC and the distillation temperature of JIS No.2 diesel fuel is between 175ºC and

354ºC (50% distillation temperature: 286ºC). In a wide IMEP range, the indicated thermal efficiency

with ordinary diesel fuel is lower than that with PRF20. This is due to the relatively larger unknown

losses with ordinary diesel fuel.

Figure 5-5 The thermal efficiency related parameters with JIS No.2 diesel fuel and PRF20 at the

optimum intake oxygen concentration conditions


59

Figure 5-6 shows the in-cylinder pressures and the rates of heat release with JIS No.2 diesel fuel and

PRF20 at the maximum rates of pressure rise of 1.0, 0.8, and 0.6 MPa/ºCA with different oxygen

concentrations at 0.6 MPa IMEP. Similar in-cylinder pressures and rates of heat release are obtained

with these two fuels with the same maximum rate of pressure rise is achieved by only changing the

intake oxygen concentration.

Figure 5-6 In-cylinder pressure and the rate of heat release (ROHR) with JIS No.2 diesel fuel and

PRF20 at the three maximum rates of pressure rise (IMEP: 0.6 MPa)

Figure 5-7 shows the intake oxygen concentrations and the thermal efficiency related parameters with

PRF20 and JIS No.2 diesel fuel at the 0.6 MPa IMEP shown with the maximum rates of pressure rise

as the abscissa. There is little difference in the degree of constant volume heat release and combustion

efficiency between these two fuels when the maximum rates of pressure rise are the same. However,

due to the larger unknown losses, the indicated thermal efficiencies with JIS No.2 diesel fuel are lower

than with PRF20. These results suggest that the larger unknown losses may be due to the fuels with

higher distillation temperature components. However, as the rates of heat release with these three fuels

are very similar as shown in Figure 5-6, and the differences in cooling loss must be very small.

Therefore, the increase in the unknown losses with JIS No.2 diesel fuel can be assumed to be mainly

due to the increased quantity of fuel which doesn’t contribute to effective combustion, the high


60

distillation components in diesel fuel which adhere to the walls of the combustion chamber without

evaporating [45,47].

Figure 5-7 Intake oxygen concentrations, O2in, and thermal efficiency related parameters with JIS

No.2 diesel fuel and PRF20 for the maximum rate of pressure rise, dp/dmax (IMEP: 0.6MPa)


61

5.4 Chapter summary

In this chapter, the effect of fuel volatility on the operational range, combustion characteristics and

emissions with premixed diesel combustion was investigated with the difference between normal

heptane, PRF20 and an ordinary diesel fuel (JIS No. 2).


1. Compared with the results between ordinary diesel fuel and normal heptane operations, the fuel

volatility has little effect on the operational range, combustion characteristics and the exhaust gas

emissions of premixed diesel combustion.

2. Although the fuel volatility has little effect on the combustion characteristics and operational

range with premixed diesel combustion, the indicated thermal efficiency with the ordinary diesel

fuel was lower than with PRF20, which have the same ignitability as the ordinary diesel fuel but

have no high distillation temperature components. This result is due to the increased quantity of

fuel which doesn’t contribute to the effective combustion, as the high distillation components in

diesel fuel adhere to the wall of the combustion chamber without evaporating.


CHAPTER 6. INFLUENCE OF FUEL DISTILLATION

TEMPERATURE ON THERMAL EFFICIENCY OF

PREMIXED DIESEL COMBUSTION


62

6.1 Chapter outline

In this experiment, to further understand the effect of fuel volatility on premixed diesel combustion

basing on the experiments in Chapter5, the influence of fuel volatility on the thermal efficiency of

premixed diesel combustion was investigated in a DI diesel engine with supercharging and a common

rail fuel injection system. Three types of ordinary diesel fuels with different distillation temperature

distributions (JIS special No.1 diesel fuel, JIS No.1 diesel fuel, and JIS special No.3 diesel fuel) and a

primary reference fuel with an octane number of 20 (PRF20) were examined to establish the influence

of fuel volatilities on the thermal efficiencies in premixed diesel combustion and in conventional

diesel combustion. An intake oxygen concentration of 11%, an equivalence ratio of 0.5 and three fuel

injection timings of 40ºCA BTDC, 25ºCA BTDC, and 10ºCA BTDC were set for the premixed diesel

combustion mode, while an intake oxygen concentration of 21%, an IMEP of 0.7 MPa, a pilot fuel

injection timing of 8ºCA BTDC, and a main injection timing of 3ºCA ATDC were used for the

conventional diesel combustion mode.

6.2 Simulation of the evaporation processes of one fuel droplet with n-heptane and n-tridecane

in the test engine condition

Figure 6-1 shows the calculated spray tip penetrations of n-heptane with the conditions of the test

engine as follows: 40ºCA BTDC (2.0 MPa, 650 K) and 25ºCA BTDC (3.5 MPa, 750 K), the injection

timings in these experiments. The calculations are based on the Hiroyasu’s equation [6].

The tbreak indicates the time of the spray breakup and the dashed horizontal lines show the distances

between the nozzle hole of injector and the cylinder liner wall at 40ºCA BTDC and the combustion

chamber wall at 25ºCA BTDC; very similar results are obtained with n-tridecane. The curves in Figure

6-1 show that the breakup time is sufficiently short and that the broken up sprays will impinge on the

walls of the cylinder liner or combustion chamber at the two injection timings. The time between the

spray breakup and the time when the spray impinges on the cylinder liner wall at 40ºCA BTDC is

around 0.6 ms, and the time between the breakup and the time when the spray impinges on the

combustion chamber wall at 25ºCA BTDC is around 0.3ms.


63

Figure 6-1 Spray tip penetrations of n-heptane in the conditions of the test engine at 40ºCA BTDC

(2.0 MPa and 650 K) and 25ºCA BTDC (3.5 MPa and 750 K)

Figure 6-2 shows the calculated results of the changes in the residual mass fraction of a fuel droplet

with n-heptane and n-tridecane at the same conditions as in Figure 6-1. A model for the evaporation of

a fuel droplet based on Spalding theory [56-57] was used to simulate the evaporation after the breakup

of the liquid fuel spray of the n-heptane and n-tridecane. The droplet is considered as a single body

with a uniform temperature, and the initial conditions in this simulation are set to an initial fuel droplet

diameter of 10 um and 313.5 K. Two in-cylinder conditions of the test engine at 40ºCA BTDC (2.0

MPa, 650 K; Figure 6-2 (a)) and 25ºCA BTDC (3.5 MPa, 750 K; Figure 6-3 (b)) with the same

pressures and temperatures as in Figure 6-1 were calculated and are shown in the figure. The transient

droplet velocity was evaluated using the equations in a reference [58]. At 40ºCA BTDC and the timing

when the spray impinges on the cylinder liner wall (0.6ms), n-heptane has evaporated while about

50% of n-tridecane still remains as liquid phase. At 25ºCA BTDC and the timing when the spray

impinges on the combustion chamber walls (0.3ms), n-heptane has evaporated while 60% of

n-tridecane still remains as liquid phase. These results suggest that with direct injection of a

low-volatility fuel including diesel fuel under the low in-cylinder gas temperature during the longer

ignition delay like in the premixed diesel combustion, wall wetting with liquid fuel spray impinging on

the combustion chamber walls is not easily avoided and the utilization of high volatility fuels would be

effective to reduce the wall wetting.


64

Figure 6-2 Changes in the residual mass fraction of a fuel droplet at (a) 40ºCA BTDC and (b) 25ºCA

BTDC (Initial droplet diameter: 10 um)


with different injection timings

Figure 6-3 shows thermal efficiency related parameters including the indicated thermal efficiency, i;

the degree of constant volume heat release,glh; the combustion efficiency calculated from the exhaust

gas components, u; and the unknown loss, w with the two diesel fuels (SDF1 and SDF3) and PRF20

under the premixed diesel combustion modes for three fuel injection timings. Here, the intake oxygen

concentration was maintained at 11% with low pressure cooled EGR, and the equivalence ratio was set

at 0.5. The ignitability of SDF1, SDF3, and PRF20 are similar while the volatilities are significantly

different as shown in Table 2-2 in Chapter 2.

As shown in Figure 6-3, while the indicated thermal efficiencies with PRF20 are higher than those

with ordinary diesel fuels (SDF1 and SDF3), the difference between SDF1 and SDF3 is small for all

three injection timings. The higher indicated thermal efficiencies with PRF20 are mainly due to lower

unknown losses than with SDF1 and SDF3 for all three injection timings. With increases in the

distillation temperature of fuels, as in the order of PRF20, SDF3, and SDF1, the unknown loss

increases at all three injection timings. Particularly, with the injection timing of 40ºCA BTDC, the

indicated thermal efficiencies with SDF1 and SDF3 are lowered, to about 35% and lower than with


65

PRF20 (above 41%), due to the larger unknown losses. The degrees of constant volume heat release

are very high for all the fuels and injection timings, but slightly lower at the late injection timing

(10ºCA BTDC). The combustion efficiency calculated from the exhaust gas components are similar

for all the fuels at one injection timing, and deteriorates significantly at the early injection timing

(40ºCA BTDC) for all three fuels. The degree of constant volume heat release and the combustion

efficiency calculated from the exhaust gas components with SDF3 are slightly lower than with SDF1

and PRF20 at the same injection timing for all three cases here. This may be due to the lower

ignitability of SDF3.


66

Figure 6-3 Thermal efficiency related parameters for special No. 1 diesel fuel, special No. 3 diesel

fuel, and PRF20 for three fuel injection timings under the premixed diesel combustion mode (Coolant

temperature: 80ºC, Intake gas temperature: 40ºC)

Figure 6-4 shows the average in-cylinder gas temperatures, the in-cylinder pressures, and the rates of

heat release with SDF1, SDF3, and PRF20 at the same conditions as in Figure 6-3. Regardless of the

injection timings and the fuels, the shape of the rate of heat release diagram features the characteristics

of premixed diesel combustion: an ignition delay substantially longer than the injection duration, the

presence of a low temperature heat release, and a single high temperature heat release. Due to the

relatively lower ignitability of SDF3, the rate of heat release with SDF3 is a little milder than with


67

SDF1 and PRF20, resulting in the slightly lower degree of constant volume heat release for all three

injection timings as mentioned previously. However, the combustion characteristics from these fuels at

one injection timing are very similar in spite of the different indicated thermal efficiencies.


(ROHR) with special No. 1 diesel fuel, special No. 3 diesel fuel, and PRF20 for three fuel injection

timings under the premixed diesel combustion mode (Coolant temperature: 80ºC, Intake gas

temperature: 40ºC)

Figure 6-5 shows the rates of heat loss to the combustion chamber walls per degree crank angle

obtained by Woschni’s formula, dQw/d, with all three fuels at the same conditions as in Figure 6-3

[59]. The rates of heat loss to the combustion chamber walls per degree crank angle with SDF1 and

SDF3 are smaller than with PRF20. Therefore, the increases in the unknown losses with the ordinary

diesel fuels are not due to the increases in heat losses to the combustion chamber walls, but due to the

increases in high distillation temperature components which adhere on the cylinder liner wall without

evaporating in the longer premixing durations. This wall-wetting causes two effects: larger unburned

fuel amounts which enter the crank case and increases in the fuel oxidized after the apparent heat

release in the cylinder, which do not contribute to the effective heat release [45,47].


68

Figure 6-5 The rate of cooling loss per degree crank angle, dQw/d, with special No.1 diesel fuel,

special No.3 diesel fuel and PRF20 for three fuel injection timings under the premixed diesel

combustion mode (Coolant temperature: 80ºC)

Figure 6-6 shows combustion related parameters including the 10% heat release crank angle, CA10,

the 50% heat release crank angle, CA50, the premixing duration, pre, and the combustion duration,

comb (duration between CA10 and CA90) with all three fuels at the same conditions as in Figure 6-3.

Here, CA10 corresponds to the ignition timing and the premixing duration, pre is defined as the crank

angle from the end of fuel injection to CA10. While CA10 and CA50 with SFD3 is slightly retarded

due to the lower ignitability and longer combustion duration with the longer premixed duration, these

parameters are not different among the fuels at the same injection timings and sufficient premixing

durations are obtained for all the cases here. An advantage CA50 as just after TDC and very short


69

combustion duration make it possible to establish the very high degree of constant volume heat release

as shown in Figure 6-3.

Figure 6-6 Combustion related parameters with special No.1 diesel fuel, special No.3 diesel fuel and

PRF20 for three fuel injection timings under the premixed diesel combustion mode (Coolant

temperature: 80ºC)

Figure 6-7 shows the exhaust gas emissions at the same conditions as in Figure 6-3. Regardless of the

injection timings and fuels, there is little difference in the CO and THC emissions for the same

injection timing. However, with the early injection timing (40ºCA BTDC), the THC and CO emissions


70

increase significantly for all three fuels. Very low NOX and smokeless combustion are realized with all

three fuel injection timings and all three fuels.

Figure 6-7 Exhaust gas emissions with special No.1 diesel fuel, special No.3 diesel fuel, and PRF20

for three fuel injection timings under the premixed diesel combustion mode (Coolant temperature:

80ºC)

6.4 Influence of fuel volatility on thermal efficiencies in premixed diesel combustion and in

conventional diesel combustion

The liquid fuel adheres on the cylinder liner or combustion chamber walls mainly during the longer

ignition delay in the premixed diesel combustion mode, and it may be assumed that there is a much

smaller influence of fuel volatility on the thermal efficiency in conventional diesel combustion mode

as the ignition occurs before the end of the fuel injection. To substantiate this, conventional diesel

combustion at an intake oxygen concentration of 21% was also examined with the ordinary diesel fuel


71

(SDF1) and PRF20. The fuel injection timing was set at a pilot injection timing of 8ºCA BTDC and a

main injection timing of 3ºCA ATDC, and the IMEP was maintained at 0.7 MPa.

Figure 6-8 shows the average in-cylinder gas temperatures, in-cylinder pressures, and the rates of heat

release with SDF1 and PRF20 under the conventional diesel combustion mode. There is little

difference in the combustion process of the two fuels.


(ROHR) with special No.1 diesel fuel and PRF20 at 0.7 IMEP for the conventional diesel combustion

mode

Figure 6-9 shows the energy balance with SDF1 and PRF20 under the premixed diesel combustion

mode (Fuel injection timing: 10ºCA BTDC) shown in Figure 6-3 and the conventional diesel

combustion mode shown in Figure 6-8. At the premixed diesel combustion mode (top graph) the

indicated work with PRF20 is larger than with SDF1 due to the decrease in the unknown loss, but at

the conventional diesel combustion mode the indicated work with the ordinary diesel fuel and PRF20

were similar and the unknown loss with PRF20 is larger than with SDF1. This is due to the very little

fuel adhering on the walls of the combustion chamber for both fuels at the conventional diesel

combustion mode, as the ignition occurs before the end of the fuel injection while there are increases

in the high distillation temperature components of the diesel fuel adhering on the walls of the

combustion chamber during the longer premixing duration between the end of the fuel injection and

the ignition timing for the premixed diesel combustion mode. Regardless of the fuels, the unburned

loss to the exhaust gas under the premixed diesel combustion mode is much larger than under the


72

conventional diesel combustion mode, which remains a problem for the premixed diesel combustion

mode even with high volatility fuels.

Figure 6-9 Energy balances with special No.1 diesel fuel and PRF20 for the conventional diesel

combustion premixed diesel combustion mode and premixed diesel combustion mode (inj = 10ºCA

BTDC)


with different coolant temperatures

Figure 6-10 shows the thermal efficiency related parameters including the indicated thermal efficiency,

i; the degree of constant volume heat release, glh; the combustion efficiency calculated from the

exhaust gas components, u; and the unknown loss, w with SDF1 and PRF20 at coolant temperatures

of 40ºC and 80ºC for three fuel injection timings. With the injection timings near TDC (θinj = 25ºCA

BTDC and 10ºCA BTDC), the indicated thermal efficiency at the high coolant temperature (80ºC) is


73

slightly higher than at 40ºC for both fuels, but the differences are not large. However, with the early

injection timing (θinj = 40ºCA BTDC) and the low coolant temperature (40ºC), the indicated thermal

efficiency with SDF1 is significantly decreased to about 29% while the deterioration with PRF20 is

still not large. This result is mainly due to the larger unknown loss with SDF1 at the lower coolant

temperature while the combustion efficiencies calculated from the exhaust gas are very similar for the

fuels when the injection timing and the coolant temperature are same.

Figure 6-10 Thermal efficiency related parameters with special No.1 diesel fuel and PRF20 for three

fuel injection timings at coolant temperatures of 80ºC and 40ºC under the premixed diesel combustion

mode

Figure 6-11 shows the average in-cylinder gas temperatures, the in-cylinder pressures, and the rates of

heat release with SDF1 and PRF20. The rates of heat release at the injection timing near top dead


74

center (θinj = 25ºCA BTDC) for both fuels do not change significantly with lowering coolant

temperatures, however the rates of heat release at the early injection timing (θinj = 40ºCA BTDC) are

retarded and become milder with lowering coolant temperatures. The changes in combustion with

lowering coolant temperatures are more significant for SDF1 than for PRF20. This is due to the

increase in the high distillation temperature components in SDF1 which adhere on the cylinder liner

wall without evaporating and do not contribute to effective combustion at the lower coolant

temperature, resulting in larger unknown losses and lower indicated thermal efficiencies.


75


(ROHR) with special No.1 diesel fuel and PRF20 for two injection timings at coolant temperatures of

80ºC and 40ºC under the premixed diesel combustion mode


76

6.6 Effect of fuel volatility on thermal efficiency of premixed diesel combustion with different

intake gas temperatures

Figure 6-12 the thermal efficiency related parameters including the indicated thermal efficiency, i;

the degree of constant volume heat release, glh; the combustion efficiency calculated from the exhaust

gas components, u; and the unknown loss, w with JIS No.1 diesel fuel (DF1) and PRF20 at intake

gas temperatures of 25ºC and 70ºC for three fuel injection timings.

With DF1 and the injection timings near top dead center (θinj = 25ºCA BTDC and 10ºCA BTDC), the

differences in the indicated thermal efficiency between the low intake gas temperature (25ºC) and the

high intake gas temperature (70ºC) are very small, but with the early injection timing (40ºCA BTDC),

the indicated thermal efficiency at the lower intake gas temperature is lower than at the higher intake

gas temperature due to the lower combustion efficiency calculated from the exhaust gas components.

With PRF20 and the injection timings near top dead center (θinj = 25ºCA BTDC and 10ºCA BTDC),

the indicated thermal efficiencies at the low intake gas temperature (25ºC) are higher than at the high

intake gas temperature (70ºC) mainly due to smaller unknown losses. This improvement with the

lower intake gas temperature may be due to the decrease in cooling loss with the lower in-cylinder gas

temperature. With PRF20 and the early injection timing (θinj = 40ºCA BTDC), the improvement in the

indicated thermal efficiency with lower intake gas temperature disappears due to the decrease in the

combustion efficiency calculated from the exhaust gas components.


77

Figure 6-12 Thermal efficiency related parameters for DF1 and PRF20 for three fuel injection timings

at two intake gas temperatures under the premixed diesel combustion mode (Coolant temperature:

80ºC)

Figure 6-13 shows the influence of the intake gas temperature on the average in-cylinder gas

temperatures, the in-cylinder pressures, and the rates of heat release with DF1 and PRF20 at two

injection timings (40ºCA BTDC and 25ºCA BTDC). With the injection timing near top dead center

(25ºCA BTDC) and the lower intake gas temperature, the heat releases are slightly retarded but the

shape of heat release does not change significantly for the two fuels. However, with the early injection

timing (40ºCA BTDC) and lower intake gas temperature (25ºC), the rate of heat release with DF1

becomes much slower while there is little slowing with PRF20. Combined with the different trends of

thermal efficiency with DF1 and PRF20 at early injection timing shown in Figure 6-12, this slower


78

heat release with DF1 at the low intake gas temperature may be due to leaner mixture formation with

larger amounts of fuel adhering on the cylinder liner wall without evaporating.


(ROHR) With No.1 diesel fuel and PRF20 for two injection timings under the premixed diesel

combustion mode (Coolant temperature: 80ºC)


79

6.7 Chapter summary

In this chapter the influence of fuel volatility on the thermal efficiency of premixed diesel combustion

with three ordinary diesel fuels with different distillation temperature distributions, and a primary

reference fuel for octane number (PRF20) was evaluated in a DI diesel engine with supercharging and

a common rail fuel injection system.


1. A model for the evaporation of a fuel droplet based on Spalding theory showed that at the

conditions of the premixed diesel combustion n-heptane spray almost all evaporates before

impinging on the cylinder liner or the combustion chamber walls while much n-tridecane spray

remains in the liquid phase at the impingement. This result suggests that with direct injection of a

low-volatility fuel including diesel fuel under the low in-cylinder gas temperature during the longer

ignition delay like in premixed diesel combustion, wall wetting with liquid fuel spray impinging on

the combustion chamber walls are not easily avoided and that the utilization of high volatility fuels

will be effective to reduce the wall wetting.

2. With the premixed diesel combustion mode, the indicated thermal efficiencies with the ordinary

diesel fuels were lower than with PRF20 although the shapes of the rate of heat release and the

combustion efficiencies calculated from the exhaust gas components were almost unchanged. With

the conventional diesel combustion mode, the indicated thermal efficiencies with the ordinary

diesel fuels and PRF20 were similar. These results suggest that the deterioration in the thermal

efficiency in the premixed diesel combustion mode with the ordinary diesel fuel is mainly due to

the adhesion of high distillation temperature components on the walls of the combustion chamber,

and occurs mainly during the premixing period (between the end of fuel injection and the ignition);

the larger amount of unburned fuel enters the crank case and the fuel oxidized after the apparent

heat release in the cylinder do not contribute to the effective heat release.

3. The deteriorations of the indicated thermal efficiencies with ordinary diesel fuels are more

significant with the early injection timing (40ºCA BTDC) than with the injection timings near TDC

(25ºCA BTDC and 10ºCA BTDC).

4. An advantage CA50 as just after TDC and very short combustion duration with premixed diesel

combustion mode will make it possible to establish a very high degree of constant volume heat

release and offers a highly attractive avenue to improve the thermal efficiency. The indicated

thermal efficiency in the premixed diesel combustion with PRF20 reaches 48%, which is much

higher than in conventional diesel combustion with diesel fuel.


80

5. With the early injection timing (40ºCA BTDC) and low coolant temperature, the indicated thermal

efficiency with diesel fuel decreases significantly while the deterioration with PRF20 is not so

large.

6. With the early injection timing (40ºCA BTDC), the indicated thermal efficiency with diesel fuel

decreases significantly at lower intake gas temperatures while the deterioration with PRF20 is not

so large.

7. The unburned loss to the exhaust gas under the premixed diesel combustion mode with PRF20 is

still much larger than under the conventional diesel combustion mode, which remains a problem for

the premixed diesel combustion mode even with high volatility fuels.


CHAPTER 7. SUMMARY


81

As a promising way to realize simultaneous reductions of NOX and smoke for meeting the stringent

emission regulations, premixed diesel combustion was paid much attention. Sufficiently long ignition

delay is required for pre-mixture preparation to avoid over-rich mixture taking part in the combustion

while the maximum pressure rise rate is suppressed to a tolerance level. Although both the control of

combustion phasing and reduction of emissions can be achieved in the premixed diesel combustion

regime with diesel fuel by late injection strategy, the operational load range is still limited by the high

pressure rise rate that is a consequence of early pre-mixed charge auto-ignition when increasing the

fuelling rate per cycle. As fuel properties have important effect on premixed diesel combustion, it may

be possible to improve the combustion process and extend the operational load range by optimization of

fuel properties.

In this study, the effects of fuel properties including ignitability, volatility, and compositions on

operational range, combustion characteristics and emissions of premixed diesel combustion were

examined in a DI diesel engine with supercharging and a common rail fuel injection system.


1. The IMEP of the premixed diesel combustion is limited by knocking with high intake oxygen

concentrations and by unstable combustion or significant increases in CO and THC emissions with

low intake oxygen concentrations regardless of the fuels examined in this experiment.

2. With increasing octane number of the primary reference fuels (PRF), the combustion phasing is

retarded, higher intake oxygen concentration for the operational range and higher optimum intake

oxygen concentrations for the thermal efficiency can be used within the tolerance limits of rapid

combustion, resulting in the higher maximum IMEP of the premixed diesel combustion and higher

indicated thermal efficiencies due to the higher combustion efficiency. Regardless of the fuels

tested in this study, the optimum intake oxygen concentration for the best thermal efficiency is

obtained near the knocking limit over a wide IMEP ranges. However, the stable operational range

of intake oxygen concentrations between the knocking and stable limits is narrower with high

octane number fuels and decreases with increases in IMEP, resulting in a reduction of combustion

control robustness.

3. The premixed diesel combustion which is obtained at lower oxygen concentrations with higher

ignitability fuels can be realized at higher intake oxygen concentrations with lower ignitability

fuels.

4. The optimum intake oxygen concentrations for the thermal efficiency and for the operational

range with normal heptane and toluene blend fuels (NTF) are higher than with the primary


82

reference fuels (PRF) with similar research octane numbers, showing the poorer ignition

characteristics than PRF .

5. The onset timings of the low temperature oxidation with normal heptane and toluene blend fuels

are retarded and the rate and quantity of the heat release with low temperature oxidation decreases

to levels lower than PRF with the same octane numbers, resulting in a more retarded and slower

main combustion. This is due to the inhibitor effect on the low temperature oxidation of toluene

and also due to the higher ignition temperature of toluene itself.


MPa IMEP both with PRF and NTF with a research octane number of 40 when the intake oxygen

concentration was optimized corresponding to IMEP and the boost pressure was set at 160 kPa

(absolute).

7. Although the fuel volatility has little effect on the combustion characteristics and operational

range with premixed diesel combustion, the indicated thermal efficiency with the ordinary diesel

fuels was lower than with PRF20, which have the same ignitability as the ordinary diesel fuels but

have no high distillation temperature components. On the other hand, with the conventional diesel

combustion mode, the indicated thermal efficiencies with the ordinary diesel fuels and PRF20

were similar. These results suggest that the deterioration in the thermal efficiency in the premixed

diesel combustion mode with the ordinary diesel fuel is mainly due to the adhesion of high

distillation temperature components on the walls of the combustion chamber, and occurs mainly

during the premixing period (between the end of fuel injection and the ignition); the larger amount

of unburned fuel enters the crank case and the fuel oxidized after the apparent heat release in the

cylinder do not contribute to the effective heat release.

8. The deteriorations of the indicated thermal efficiencies with ordinary diesel fuels are more

significant with the early injection timing (40ºCA BTDC) than with the injection timings near

TDC (25ºCA BTDC and 10ºCA BTDC).

9. An advantage CA50 as just after TDC and very short combustion duration with premixed diesel

combustion mode will make it possible to establish a very high degree of constant volume heat

release and offers a highly attractive avenue to improve the thermal efficiency. The indicated

thermal efficiency in the premixed diesel combustion with PRF20 reaches 48%, which is much

higher than in conventional diesel combustion with diesel fuel.

10. With the early injection timing (40ºCA BTDC) and low coolant temperature or lower intake gas

temperatures, the indicated thermal efficiency with diesel fuel decreases significantly while the


83

deterioration with PRF20 is not so large.

11. The unburned loss to the exhaust gas under the premixed diesel combustion mode with PRF20 is

still much larger than under the conventional diesel combustion mode, which remains a problem

for the premixed diesel combustion mode even with high volatility fuels.


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ACKNOWLEDGEMENTS

I am indebted to express my gratitude towards, Professor Hideyuki Ogawa to accept me as his student,

to give me opportunity to work on influence of fuel properties on premixed diesel combustion, and to

guide me with patience throughout my research works at Hokkaido University.

I also would like to acknowledge to Associate Professor Gen Shibata, Associate Professor Tie Li for

their valuable suggestions during performing this research. Thanks go to Professor Takeyuki

Kamimoto for his advice on my paper accepted by International Journal of Engine Research.

I would like to acknowledge to Associate Professor Gen Shibata, Professor Takemi Chikahisa and

Prof Osamu Fujita for their efforts to review my thesis and to give valuable suggestions during my

close defense.

Special thanks go to Mr. Yamazaki for his help and support on experimental work and also share

valuable experience of his expertise.

I would like to acknowledge to the China Scholarship Council (CSC) for awarding scholarship during

doctoral program at Hokkaido University.

I would like to acknowledge to International Student Center at Hokkaido University for providing me

the chance to study Japanese during doctoral program.

I would also like to thank all of members of Applied Thermal Engineering laboratory; Special thanks

go to Mr. Inaba, Mr. Obe, and Mr. Sakane for their support and help during experiments.

Finally, I wish to thank my parents, my sister and friends for their support during my stay in Japan. In

particular I thank my wife, Jing Qu for her love, understanding and support during my stay in Japan.

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