FACULTY OF ENGINEERING & TECHNOLOGY (AUTOMOBILE ...shodhganga.inflibnet.ac.in/bitstream/10603/25/1/venkateshbabu-all-chapters.pdf6.20 Variation of cylinder gas temperature with crank

STUDIES ON THE ENHANCEMENT OF DIESEL ENGINE COMBUSTION THROUGH

THE USE OF FUEL ADDITIVES AND IN-CYLINDER TURBULENCE INDUCEMENT TECHNIQUES

A THESIS

Submitted by

R. VENKATESH BABU

[Reg.No. D04AM001]

In fulfillment for the award of the degree

of

DOCTOR OF PHILOSOPHY

FACULTY OF ENGINEERING & TECHNOLOGY (AUTOMOBILE ENGINEERING)

BHARATH UNIVERSITY CHENNAI- 600 073, INDIA.

OCTOBER - 2008

i

ACKNOWLEDGEMENT

I would like to express my sincere gratitude to Dr.S.Jagathrakshagan, Honorable

Chancellor, Bharath University for his encouragement and motivation that propelled me

to submit the First Thesis of the University.

I thank Dr.S.Sendilvelan, my guide for his able guidance and suggestions that

helped me in achieving the aim of this research work.

I would like to put on record the kindness shown by Shri.J.Sundeep Aanand,

Pro-Chancellor and Mrs.Shwetha Sundeep Anand by granting me concession in the

tuition fees.

I also thank Prof.Dr.K.P.Thooyamani, Vice-Chancellor, Bharath University for

his kind words and advices that went a long way in the completion of this research. I also

thank Prof.Dr.M.P.Chockalingam, Dean, R&D and Prof.M.Prem Jeya Kumar, HOD,

Dept. of Automobile Engineering, Bharath University for permitting me to use the

Automobile Engineering Laboratory of the University.

I would also like to thank all the teaching and non teaching staff of the Dept. of

Automobile Engineering for their kind cooperation. I also thank all the members of the

Doctoral Committee for their kind suggestions which helped me overcome many hurdles

in the course of this work.

I thank my family, friends and all my well wishers whose good wishes have

brought me where Iam and would continue to take me forward in the right direction.

ii

ABSTRACT

Key Words: Direct Injection Diesel Engines, In-Cylinder Turbulence, Bluff Bodies,

Internal Jets, Fuel Additives, Emission.

Owing to direct injection (DI) diesel engines becoming acceptable choice as

prime movers in many applications, it has become imperative to improve their fuel

consumption and emission characteristics. For this purpose, several attempts are ongoing

to improve the emission characteristics of DI diesel engines without having to sacrifice

their fuel consumption advantage. In diesel engines, fuel is injected near compression top

dead center and hence the requirements of fuel-air mixing are quite stringent. The fuel-air

mixing process therefore remains at the core of the diesel engine combustion and

emission problems. Beside better fuel-air mixing, the improvements in diesel engines are

also possible through changes in fuel.

In the present work, the changes in fuel and fuel-air mixing process are tested

independently and together to improve diesel engine performance and emission

characteristics. A simpler method of producing in-cylinder turbulence has been arrived at

and investigated. The same has been combined with the best of the additives tested to

attain a significant decrease in fuel consumption and exhaust smoke concentration.

Six polymer based additives of varying properties are mixed in different

proportions in diesel fuel and their experimental results compared. Beside, comparing the

measured performance and the exhaust emissions (exhaust smoke and oxides of nitrogen)

of various fuel-additive combinations, a detailed and systematic combustion analysis of

the acquired cylinder pressure histories on these samples has been attempted for

understanding the effect of the additives on the engine combustion characteristics. From

iii

this analysis, it is observed that the engine combustion has become smoother in presence

of certain additives proportions used here. The maximum improvements in BSFC and

exhaust smoke level are found to be 13 % and 37.5 % respectively in case of Additive 6

(at 2% by volume) fuel-additive combinations as compared to that of the base diesel fuel.

The optimum choice in terms of additive cost is found to be Additive 1 (at 0.5% by

volume) with an improvement of 7.6 % in BSFC and 36.8% in the exhaust smoke level.

In the second stage of the work, the effects of inducing in-cylinder turbulence

through bluff bodies or internal jets on the diesel engine are investigated. In the work

carried out here, the bluff bodies are placed horizontally across the piston cavity in the

form of rods or rods wound with thin wire in different orientations with respect to the

piston pin axis. The jet turbulence is introduced by holes on the piston crown, allowing a

tangential entry of the working fluid into the piston cavity along the direction of swirl.

The effect of size, position and number of jets has been investigated. In general,

horizontal bluff bodies do not result in significant advantage in fuel economy and smoke

levels, but some reduction in NOx concentration is observed. More importantly, it is

observed that the internal jets introduced through the tangential holes showed

improvement in the engine brake thermal efficiency and exhaust smoke level with only a

marginal increase in NOx concentration.

An attempt to predict mixing effects of internal jets through an available

commercial CFD package STAR-CD reveals that the turbulent kinetic energy and the

eddy dissipation rate is maximum in the case of the internal jets with 3 mm diameter. The

experimental results concerning performance, combustion and emissions of the engine

corroborates with this mixing predictions.

iv

Finally, a representative study on the combined effects of the best fuel additive

combination and the best internal jet configuration suggests superior performance in

terms of fuel consumption and even better exhaust smoke emission to the independent

changes. In the combined case there is a decrease of 9.8% in BSFC and about 38.5%

decrease in the exhaust smoke level with a marginal increase of about 4.5% in of NO

level vis-à-vis the base engine.

v

TABLE OF CONTENTS

ACKNOWLEDGMENT ………………………………………………………. i

ABSTRACT ……………………………………………………………………. ii

LIST OF TABLES …………………………………………………………….. vi

LIST OF FIGURES …………………………………………………………… vii

ABBREVIATIONS ……………………………………………………………. Xi

CHAPTER 1 INTRODUCTION ………………………………………. 1

CHAPTER 2 LITERATURE SURVEY ………………………………. 5

2.1 Additive with fuel …………………………………………………... 5

2.2 In-cylinder turbulence inducement …………………………………. 8

2.2.1 Importance of fuel-air mixing ……………………………………… 8

2.3 Closure ……………………………………………………………… 22

CHAPTER 3 OBJECTIVE OF THE PRESENT WORK ……………. 23

3.1 Motivation …………………………………………………………… 23

3.2 Objective ……………………………………………………………... 23

CHAPTER 4 EXPERIMENTAL WORK ……………………………... 25

4.1 Test engine …………………………………………………………… 25

4.2 Engine instrumentation ………………………………………………. 27

4.2.1 Pressure measurement ……………………………………………….. 27

4.2.2 TDC encoder ………………………………………………………… 27

4.2.3 Analog to Digital converter ………………………………………….. 28

4.2.4 Power measurement ……………………………….............................. 28

4.2.5 Fuel flow rate measurement …………………………………………. 28

4.2.6 Air flow rate measurement …………………………………………... 29

4.2.7 Temperature measurement …………………………………………... 29

4.2.8 Smoke measurement …………………………………………………. 29

4.2.9 Measurement of oxides of nitrogen ………………………………….. 30

4.3 Engine experimentation ……………………………………………… 30

4.3.1 Experiments with fuel modifications using additives ……………….. 31

4.3.2 Experiments with engine modifications ……………………………... 31

4.3.2.1 Turbulence inducement through bluff bodies ………………………... 32

4.3.2.2 Internal jets …………………………………………………………... 33

CHAPTER 5 ANALYSIS PROCEDURE 35

vi

5.1 Combustion analysis …………………………………………………. 37

5.2 Mixing / Turbulence analysis ………………………………………... 38

CHAPTER 6 RESULTS AND DISCUSSION 40

6.1 Additive with fuel ……………………………………………………. 40

6.2 In-cylinder turbulence ……………………………………………….. 83

6.2.1 Effect of bluff bodies ………………………………………………… 84

6.2.2 Effect of internal jets ………………………………………………… 85

6.3 Parametric studies ……………………………………………………. 86

6.3.1 Effect of number and position of the internal jets …………………… 89

6.3.2 Effect of size of the internal jets ……………………………………... 90

6.4 Combined effects of fuel additive and in-cylinder turbulence

modifications …………………………………………………………

99

CHAPTER 7 CONCLUSIONS AND SCOPE FOR FUTURE WORK 107

7.1 Additive with fuel ……………………………………………………. 107

7.2 In-cylinder turbulence inducement …………………………………... 108

7.3 Scope for future work ………………………………………………... 109

REFERENCES …………………………………………………………….…... 110

LIST OF PAPERS SUBMITTED ON THE BASIS OF THIS THESIS ….... 115

LIST OF TABLES

Table Title Page

4.1 Engine specifications ……………………………………………………. 26

4.2 Dimensions of the elements used for generating in cylinder turbulence .. 33

4.3 Several configurations of internal jets …………………………………... 34

6.1 Range of variations for different performance parameters at full load …. 41

6.2 Various configurations for in-cylinder turbulence inducement ………… 84

vii

LIST OF FIGURES

Fig. Title Page

2.1 Precombustion chamber (Maleev, 1987) ……………………………….. 10

2.2 Turbulence chamber (Maleev, 1987) …………………………………… 11

2.3 Energy cell combustion chamber (Maleev, 1987) ……………………… 12

2.4 Different combustion cavity shapes (Shigemori et al. 1983) …………… 13

2.5 Combustion chamber geometry (Montjir et al., 2000) ………………….. 15

2.6 DS combustion chamber (Rong et al., 2000) …………………………… 15

2.7 Configuration of test engine with an air cell (Kamimoto et al., 1983) …. 16

2.8 Concept of MIW head for the NICS-MH engine (Lin et al., 1995) …….. 17

2.9 MIW head used in the experiments (Lin et al., 1995) …………………... 17

2.10 Four combustion chambers used in the experiments (Lin et al., 1995) … 17

2.11 Fuel spray location of MULDIC (Hashizume et al., 1998) ……………... 18

2.12 Cross section of the CCD system (Konno et al., 1992) …………………. 19

2.13 Cylinder head configuration (Choi et al., 1995) ………………………… 20

4.1 Schematic of Experimental Setup ………………………………………. 26

4.2 Arrangement of bluff bodies on different orientations ………………….. 32

5.1 Slider crank mechanism of an IC engine ……………………………….. 35

5.2 Pressure volume diagram of a thermodynamic cycle …………………… 36

6.1 Variation of BSFC with load (additive - 1) ……………………………... 42






6.7 Variation of smoke with load (additive - 1) …………………………….. 45






6.13 Variation of NO with load (additive - 1) ………………………………... 49



viii

6.16 Variation of NO with load (additive - 4) ……………………………….. 50



6.19 Variation of cylinder gas temperature with crank angle (additive - 1) …. 52






6.25 Variation of IMEP with load (additive - 1) ……………………………... 55






6.31 Variation of peak cylinder temperature with load (additive-1) …………. 59






6.37 Variation of total combustion duration with load (additive-1) …………. 62






6.43 Variation of exhaust gas temperature with load (additive-1) …………… 65






6.49 Variation of cylinder wall temperature with load (additive-1) ………… 68


ix





6.55 Variation of BSFC with percentage additive (additive- 1) ……………... 71

6.56 Variation of BSFC with percentage additive (additive-2) …………….... 71

6.57 Variation of BSFC with percentage additive (additive-3) ……………... 72




6.61 Variation of smoke with percentage additive (additive-1) ……………... 74






6.67 Variation of brake specific fuel consumption with load ………………... 80

6.68 Variation of peak pressure with load ……………………………………. 80

6.69 Variation of maximum rate of pressure rise with load ………………….. 81

6.70 Variation of ignition delay with load …………………………………… 81

6.71 Variation of smoke number with load …………………………………... 82

6.72 Variation of nitric oxide with load ……………………………………… 82

6.73 Ignition delay for horizontal rods with and without wire ……………….. 86

6.74 Peak pressure for horizontal rods with and without wire, two internal jet

along with base engine ………………………………………………….. 86

6.75 BSFC, smoke level and NOx for horizontal rods with and without wire,

two internal jet along with base engine …………………………………. 87

6.76 Variation of brake thermal efficiency with load ………………………... 91




6.80 Variation of combustion duration with load ……………………………. 93


6.82 Variation of nitric oxide with load ……………………………………… 94

6.83 Variation of brake thermal efficiency with load ………………………... 94

x




6.87 Variation of combustion duration with load ……………………………. 96


6.89 Variation of nitric oxide with load …………………………………….... 97

6.90 Comparison of the brake specific fuel consumption of the best of

additive, internal jet and the combined case with the base engine ……… 99

6.91 Comparison of the peak pressure of the best of additive internal jets and

the combined case with the base engine ………………………………… 99

6.92 Comparison of the max rate of pressure rise of the best of additive

internal jets and the combined case with the base engine ………………. 100

6.93 Comparison of the ignition delay of the best of additive internal jets and


6.94 Comparison of the combustion duration of the best of additive internal

jets and the combined case with the base engine ……………………….. 101

6.95 Comparison of the smoke number of the best of additive internal jets

and the combined case with the base engine ……………………………. 101

6.96 Comparison of the Nitric oxide of the best of additive, internal jets and


xi

ABBREVIATIONS

A Late Stage Injection

A/F Air Fuel Ratio

ATDC After Top Dead Center

BSFC Brake Specific Fuel Consumption, g/kWh

BTH Brake Thermal Efficiency, %

CD Combustion Duration (OCA)

CI Compression Ignition

CO Carbon monoxide

CWT Cylinder Wall Temperature, °C

DI Direct Injection

E Early Stage Injection

EGR Exhaust gas recirculation

EGT Exhaust gas temperature, °C

HBP High Back Pressure

HC Hydrocarbon

ID Ignition Delay, COCA)

IDI Indirect Injection

IMEP Indicated Mean Effective Pressure, bar

IP Indicated Power, kW

M Main Injection

MRPR Maximum Rate of Pressure Rise (bar/ca)

NO Nitric Oxide

NOx Oxides of Nitrogen

xii

P Pilot Injection

Pddmax Peak Second Rate Of Change Of Cylinder Pressure, bars/s2

Pdmax Peak First Rate Of Change Of Cylinder Pressure, bars/s

Pmax Peak Cylinder Pressure, bar

RNG Re-Normalized Group

SI Spark Ignition

SMK Bosch Smoke Unit, (BSU)

TCD Total Combustion Duration (°CA)

TDC Top Dead Center

TOC Calculated Peak Cylinder Temperature, K

For Legends in Additive Graphs

Each additive test is designated with a letter a followed by six or seven numerical digits

where the first digit after 'a' denotes the additive number and the next digits on division

by 10 represent the percentage additive in fuel sample. Last two or three digits represent

the percentage load pertaining to the data. For example, 120100 means data involves

additive number 1 at 2 % addition (i.e. 20 ml additive added to 1 liter base diesel) tested

at l00 percent engine load.

1

CHAPTER 1

INTRODUCTION

Diesel engines, particularly direct injection types, have been an important choice

as prime movers in heavy-duty applications such as on-road, off-road, marine and

industrial usage due to their high brake thermal efficiency. In diesel engines, a high

cetane fuel is injected into the cylinder and mixed with air. The fuel-air mixture thus

formed bums under compression ignition. Diesel engine processes exhibit complex

features, perhaps more than any other mechanical device. Despite these complexities,

diesel engines have gone through very ambitious developments over last one century or

so, and still the margin for their improvement are relatively wide. The improved

efficiency in diesel engines is caused by the relatively high compression ratios, low

pumping losses due to unthrottled mode of operation, the use of lean mixtures, and the

fact that crevice volumes have air or products of combustion instead of unburned fuel

mixture. Diesel engines are basically low speed high torque engines, suitable for hauling

loads in trucks. They have high backup torque unlike gasoline engines and thus

eliminating need of frequent gear changes when used in automobile applications. Diesel

engines are sturdier and withstand rough duties. Their power rating is limited only by

smoke and not the maximum power output delivered.

As far fuels in diesel engines are concerned, the alternative fuels such as biomass,

vegetable oil, alcohol, hydrogen, liquefied petroleum gas, compressed natural gas etc. are

being used them in straight or dual fuel modes without many problems. With the

increasing concern about the green house effects on the world climate, lower CO2

emission of diesel engine (about 30%) compared to gasoline engine, remains an

advantage. The suitability of diesel engine for supercharging, which is extensively used

2

on stationary and mobile applications, leads to a high power output and reduced smoke

and other exhaust emissions from this variety of engine.

From the stand point of their disadvantages, the diesel engines emit high oxides of

nitrogen, smoke and particulate emissions in exhaust. Larger forces arising out of high

compression ratio on various parts of the engine makes these engines heavier. Also, due

to lean mixture operation, their power to weight ratio and the power to volume ratio are

lower than the SI engine. Due to heterogeneous nature of charge, there is no regular flame

propagation like in SI engine, hence multiple auto ignition mode makes CI engines much

more noisier than SI engines. A higher ignition delay in diesel engine leads to a greater

accumulation of fuel prior to the onset of combustion causing a higher rate of pressure

rise and consequently the roughness in engine operation.

The fuel economy and exhaust emission regulations, new technologies,

development time and cost reduction require increasingly sophisticated solutions to

improve the diesel engine performance and reduce exhaust emissions. Combustion

process is central to the majority of engine development related issues and requires varied

approaches to achieve desired improvements. The diesel engine combustion process

involves flows of air and fuel into the combustion chamber, their mixing and ignition.

The degree of homogeneity of the air-fuel mixture, cycle-to-cycle variation of

thermodynamic and mixing parameters, -and turbulence intensity variations are the

conditions affecting engine performance and emission characteristics.

Several methods available for improving diesel engine performance and emissions,

namely oxides of nitrogen, smoke and particulates, include high pressure injection, split

injection, water injection, exhaust gas recirculation, water diesel emulsion, retarded

injection timing, intake charge oxygen enrichment and combustion chamber design for

3

better fuel-air mixing. Among these methods, some require modifications in fuel injection

system, while many other methods include modifications in the combustion chamber or

fuel. This work, however, concerns investigating the effects of the modifications in the

engine combustion chamber and the fuel in order to achieve improvements in diesel

engine performance and emission characteristics. The increase in demand for petroleum

fuels and consequent depletion of their reserves has given rise to a need for identifying

and investigating new energy resources and/or finding the optimum way of using the

present resources. In this regard, generally the following two approaches are pursued

a. Tailoring fuel at the refining stage i.e. improving refining processes for producing

better quality fuel from different crude oils, and

b. Improving performance of available fuel i.e. using some additives for improving the

quality of existing fuels to a desired level.

The effects of fuel quality variations on diesel engine emissions is rather complex

due to wide variation of engine response to fuel quality changes and the extent of inter-

correlation of the various fuel variables.

The diesel fuel has higher carbon content and is heavier than other conventional

fuels and thus poses problems during use in engine. Due to its high freezing point, diesel

fuel causes blockage of filters and nozzles especially under cold conditions. Towards

these and other problems, the use of additive is in vogue. Some additives achieve a

specific objective of improving either physical or chemical characteristics of the fuel or

improving the combustion characteristics.

In diesel engines, the fine atomized fuel particles sprayed into the cylinder mix

with air during compression stroke. For efficient combustion in diesel engines, the fuel

and air are required to attain proper mixing between them. The requirements of in-

4

cylinder fuel-air mixing in desired range of quality (proper fuel-air mixture), has to be

supported by organized and unorganized in-cylinder air motion such as swirl, turbulence,

etc. There are various techniques used to generate turbulence in engine combustion

chamber, involving either hardware modifications, or using process like pre-combustion.

Also, fuel injection in finely atomized form produces turbulence.

In order to provide complete combustion at a constant rate, there is common

design objective of bringing sufficient air in contact with the injected fuel particles. For

this purpose, the piston crown and the cylinder head are shaped to induce a swirling

motion to air, while during compression piston is moving towards TDC. The production

of turbulence by different means, however, is considered necessary for better fuel-air

mixing. The complexities of production and the higher costs of these methods of creating

turbulence are the limiting factors in their wider use.

The present work is aimed at studying the effects of modifications in fuel and fuel-

air mixing respectively for improving diesel engine combustion and emission

characteristics. These modifications include use of

a. Polymer based additives in fuel, and

b. In-cylinder turbulence inducement through bluff bodies or internal jets.

The discussions in this thesis are focused mainly on these two aspects concerning

fuel with additives and the turbulence inducement for better fuel-air mixing. A discussion

on the existing literature concerning these aspects is presented in the next chapter, prior to

the details of the work carried out during this investigation.

5

CHAPTER 2

LITERATURE SURVEY

The literature survey regarding use of additive with fuel in-cylinder turbulence

inducement aspect investigated are reviewed and discussed in this chapter.

2.1 ADDITIVE WITH FUEL

The increase in demand for petroleum fuels and consequent depletion of their

reserves has given rise to the need for investigating new energy resources or finding the

optimum way of using the present resources. In this regard, two approaches are pursued

a) Improving refining processes for producing better quality fuel from different crude

oils, that is, tailoring fuel at the refining stage, and

b) Using some additives for improving the quality of existing fuels to a desired level,

which is, improving performance of available fuel.

The effects of fuel quality variations on diesel engine emissions is complicated by

the wide variation of the engine response to the fuel quality changes and the extent of

inter-correlation of the various fuel variables. In engine literature, many investigators

have reported. Betroli et al. (1993) suggest that the particulate emission reduction could

be attained using the ash less additive technology. The different fuel characteristics are

given in Table 2.1. They found that it is necessary to use a conditioning period prior to

emission tests.

Kouremenous et al. (1999) examined the effect of the fuel composition and

physical properties on the mechanism of combustion and pollutant formation. A number

of fuels having different density, viscosity, chemical composition, (especially aromatics

type), are used in their investigation and found that the fuel properties namely density and

viscosity are more important than fuel composition (aromatics) in respect of engine

6

performance and emissions. The total aromatic content, however, has more influence on

engine performance and emissions rather than the individual aromatics.

Hajdukovic et al. (2000) reported that the toxicity of diesel fuel is generally

attributed to soluble aromatic compounds. Alkyl derivatives of benzene and polycyclic

aromatic hydrocarbons are considered as most harmful. New oxygen and nitrogen

derivatives of hydrocarbons are formed as a result of oxidative and pyrolytic processes

during combustion.

The diesel fuel being heavier and having higher carbon content has some

problems when used in an engine. Due to its high freezing point, it is known to cause

blockage of filters and nozzles especially under cold conditions. The routine use of fuel

additive in diesel began in 1960's in Europe as cold flow improvers. The additives added

in parts per millions (ppm) levels achieve a specific objective of either improving the

physical or chemical characteristics of the fuel or improving the combustion

characteristics. There are many other functions of additives. Based on the function and

additive concept, they are reported to be classified (Owen Kieth et al, 1990) as

antioxidants and stabilizers, metal deactivators, cetane improvers, combustion improvers,

detergents, corrosion inhibitors, anti static additives, dehazers and demulsifiers, anti-icers,

biocides, anti-foamants, odor masks and odorants, dyers and markers and drag reducers.

Kidoguchi et al. (2000) in their investigations reported that in fuels with higher

aromatics content, the pyrolysis of fuel will not be satisfactory and therefore there are

local high temperature regions on account of higher adiabatic flame temperature

capability of ring structure hydrocarbons. The aromatic compounds are very compact

with very less surface to volume ratio compared to long chain normal polymers. They

have higher C/H ratio and also cm ratio per unit volume. They are also more reactive

because of lower C-C bond strength compared to C-H bonds.

7

Hence, in the absence of air, they are prone to higher cracking, pyrolysis and

agglomeration with other aromatic molecules nearby during the initial stages of

combustion. Their adiabatic flame temperatures are also very high and as a result, soot

formation increases (Hirao et al., 1988). Due to higher bond strength of O-H bonds

compared to C-H and C-C bonds, O-H bonds break up in presence of high local

temperatures and bring the local temperatures down. This decreases the possibility of

formation of NOx. The O-H bonds are reformed as the temperatures decrease and the

absorbed energy is given back.

Jensen et al (1983) observed that the concentrations of alkyl homologues of PAH

and oxy-PAH in the particulates were found to decrease with increasing cylinder exhaust

temperatures. The degree of alkylation for the most abundant homologue of these

compounds increased by one to two carbons as the cylinder exhaust temperature

decreased. The inverse relationship between engine temperature and production of

extractable organics suggests one possible emission control strategy. The post combustion

reactor might achieve reduction of PM associated with organics. To evaluate the

feasibility of such an engine modification, both particulate and vapour emissions need to

be collected simultaneously. This will allow proper correlation of particulate vapour with

the engine conditions. Alkyl homologue analysis of diesel emissions provides information

which may lead to selection of engine operating conditions that will reduce the

environmental impact of diesel emissions.

It is reported that

a) Iso-propyl nitrate reduces both aldehyde and CO level without much effect on NOx.

b) Iso and Iso-amyl nitrate and di-tertiary butyl peroxide reduce NOx by generating

alkoxyl radicals.

8

Stage de Caro et al. (2001) studied the effect of two organic additives for their

properties and to investigate their effect on diesel - ethanol mixture they tested them in

the DI and IDI engines. Additives bring stability to the diesel ethanol mixture. Cetane

number decreased in the presence of alcohol and also the dynamic viscosity, and heat

content increases the volatility. Diesel / ethanol blends with low ethanol content have

little effect on the contents of the pollutant gases from the indirect injection engines

whereas a reduction is observed with DI engines. DI engines are more sensitive than IDI

engines to the fuel cetane number. Adding ethanol leads to a reduction in the smoke and

particulates levels emitted in the exhaust. In the presence of additive, the cycle-to-cycle

variation of IMEP was reduced.

Kulinowski et al. (1993) in his review suggested that diesel fuel additives such as

cetane improvers, combustion improvers, diesel detergents, low aromatic and sulphur

content in fuel and lubricity additives can give a desirable effect. They concluded that a

properly formulated diesel additive with the above measures will result in desirable

changes in the emissions and performance of the engine.

2.2 IN-CYLINDER TURBULENCE INDUCEMENT

2.2.1 Importance of fuel-air mixing

The role of air motion in diesel engines is well recognized for the purpose of fuel-

air mixing which is central to the engine combustion and emission characteristics. The

effect of organized or unorganized air motion in the engine combustion chamber is

generally considered responsible to set in a particular flow field influencing the fuel- air

mixing pattern. In general, the air motion responsible for mixing can be considered to

affect the in-cylinder turbulence prior to the fuel-air mixing. The generation of in-cylinder

9

turbulence has been a widely investigated aspect in the context of internal combustion

engine context, particularly diesel engines, where mixing process assumes primary

importance. From the vast literature that exists in this area, the present discussion is so

organized that the state of the art concerning generation or inducement of turbulence is

generally covered.

These aspects include:

a) Combustion System

b) Combustion Chamber geometry shapes

i) Piston cavity

ii) Cylinder head

c) Injection process

i) High pressure injection

ii) Auxiliary gas / Air injection

d) Bluff bodies

e) CFD analysis for turbulence

a) Combustion System

In diesel engines, fuel is injected and mixed with hot and compressed air in the

cylinder. The presence of air movement generally termed as turbulence is considered

necessary to enhance fuel-air mixing for better combustion. There are several techniques

used for creating turbulence in the engine. These techniques use either processes like

injection, precombustion etc., or hardware modifications such as air cells etc. Fuel is

distributed in the cylinders of a diesel engine by injection nozzles, which atomize the fuel

and direct it to the desired portions of the combustion space.

10

Fuel injection itself creates some turbulence, but not enough for efficient

combustion. This conditioning, called pre-combustion, involves a partial burning of the

fuel before it enters the main combustion space. Precombustion helps to create the

turbulence needed for the fuel and air to be properly mix. Because of differences in

designs, the manner in which precombustion aids in creating turbulence differs from one

type of auxiliary combustion chamber to another. A spherical precombustion chamber is

shown in Fig. 2.1. The precombustion chamber is located in the cylinder head and is

connected to the main combustion space of the cylinder by a multiple orifice called a

burner. During the compression event, a relatively small volume of compression-heated

air is forced through the burner into the precombustion chamber. Heat stored by the

burner increases the temperature of the compressed air and facilitates initial ignition. Fuel

is atomized and sprayed into the hot air in the precombustion chamber and combustion

begins. Only a small part of the fuel is burned in the precombustion chamber because of

the limited amount of oxygen.

Fig. 2.1. Precombustion chamber (Maleev, 1987)

The fuel that does burn in the chamber creates enough heat and pressure to force

the fuel, as injection continues, into the cylinder at higher velocity. The velocity of the

fuel entering the main combustion space and the shape of the piston crown help creating

11

the necessary turbulence within the cylinder. Engines that have precombustion chambers

do not require high fuel injection pressures as great as engines that have open-type

configurations. Also, the spray of injected fuel can be coarser, since the precombustion

chamber functions to atomize the fuel further before the fuel enters the cylinder. The

engines have auxiliary combustion chambers, which differ from precombustion chambers

such that almost all of the air supplied to the cylinder during the intake event is forced

into the auxiliary chamber during the compression stroke. Auxiliary chambers in which

this occurs .are sometimes referred to as Turbulence chambers as shown in Fig 2.2.

Fig. 2.2. Turbulence chamber (Maleev, 1987)

The turbulence is created in the auxiliary chamber in compression, injection and

combustion periods. In engines with turbulence chambers, there is very little clearance

between the top of the piston and the head when the piston reaches TDC. For this reason,

a high percentage of the air in the cylinder is forced into the turbulence chamber during

the compression event. The shape of the chamber (usually spherical) and the size of the

opening through which the air must pass help to create turbulence.

12

The Lanova cell is the energy cell divided chamber type. Fig. 2.3 shows cross-

sectional top and side views of a divided auxiliary combustion chamber. This design

employs a combustion chamber consisting of two rounded spaces cast in the cylinder

head. The inlet and exhaust valves open into the main combustion chamber. The fuel-

injection nozzle lies horizontally pointing across the narrow section where the lobes join.

Fig.2.3 Energy cell combustion chamber (Maleev, 1987)

Opposite to the nozzle is the two-part energy cell, which contains less than 20

percent of the main-chamber volume. During the compression stroke, the piston forces air

into the energy cell. Near the end of the stroke, the nozzle sprays fuel across the main

chamber in the direction of the mouth of the energy cell. While the fuel charge is

traveling across the center of the main chamber, between a third and a half of the fuel

mixes with the hot air and bums at once. The remainder of the fuel enters the energy cell

and starts to bum there, being ignited from the fuel already burning in the main chamber.

At this point, the cell pressure rises sharply, causing the products of combustion to flow at

13

high velocity back into the main combustion space. This sets up a rapid swirling

movement of fuel and air in each lobe of the main chamber, promoting the final fuel-air

mixing and ensuring complete combustion. The two restricted openings of the energy cell

control the time and rate of expulsion of the turbulence-creating blast from the energy cell

into the main combustion space. Therefore, the rate of pressure rise on the piston is

gradual, resulting in smooth engine operation. However, turbulence in a divided

combustion chamber is dependent on thermal expansion caused by combustion in the

energy cell and not on engine speed as in other types of auxiliary combustion chambers.

b. Combustion Chamber geometry shapes

i) Combustion chambers having cavity in piston

Shigemori et al. (1983) developed a combustion chamber (refer Fig. 2.4) with

turbulence induced intake port and optimum fuel injection equipment. They reported that

the HMMS-III has the superior performance with a 3 mm nozzle protrusion at all speeds

due to short combustion period & active reactions in the second stage of combustion.

Fig. 2.4 Different combustion cavity shapes (Shigemori et al. 1983)

Saito et al (1986) investigated the effect of the combustion geometry on

combustion with special emphasis focused on the re-entrant combustion chamber. They

compared the conventional combustion chambers and the reentrant in terms of

14

combustion process, engine performance and NOx and smoke emissions. They found that

the reentrant chamber reduces ignition lag and provides better fuel economy with delayed

injection timing, which is attributed to the effect produced by the hotter surface of the re-

entrant chamber. Also combustion is enhanced with reduced smoke emission due to

higher velocities induced around TDC accompanying much turbulence.

The combustion chamber geometry, the shape of the cavity entrance, bottom

comer radius and the position where spray impinges on the wall were varied to investigate

their effects on the spray development in the chamber using a common rail injection

system (refer Fig. 2.5). In this they have studied the experiments with the focus on the

following parameters, that is, the spray spreading area, equivalent wall jet diameter and

spray path. They found that the reentrant cavity with round lip produces larger spray

volumes and wider spray spreading. For effect on impinging position they stated that the

fuel impingement just on the lip comer produces the maximum spreading area. They also

concluded that introduction of a bottom comer radius helps to disperse the fuel

accumulated at the bottom comer and the spray volume increases.

Rong et al. (2000) developed new combustion system (DSCS) Double Swirls

combustion System (DSCS) as shown in Fig.2.6. This combustion chamber is made of

two dishes, smaller in the middle of the bigger one. They reported to have reduced fuel

consumption by 5-10%. This is attributed to the fact that the fuel jets collides with the

ridges of the DSCS combustion chamber and then splits and form double swirl, hence

mixing and burning are efficiently carried out.

15

Fig. 2.5 Combustion chamber geometry (Montjir et al., 2000)

Fig. 2.6. DS combustion chamber (Rong et al., 2000)

ii) Combustion chamber having cavity in Cylinder head

Kamimoto et al (1983) studied the effect of air cell fitted on the cylinder head for

soot reduction in a DI diesel engine. The air cell fitted engine is as shown in Fig. 2.7. Air

is accumulated in the air cell during compression stroke and is injected into the main

chamber during the period after the end of the injection. At this instant the air jet stirs the

stagnant flame and promotes soot oxidation. They found that the soot emission was lower

by 30% in the higher load operation than that of the conventional type of engine. NO

concentration is lower in case of air cell system. The air cell fitted engine has higher

16

specific fuel consumption at low load condition because there is loss in the effective

work, which is the air movement between the combustion chamber and the air cell.

Fig. 2.7 Configuration of test engine with an air cell (Kamimoto et al., 1983)

Lin et al (1995) in their investigation designed a multi-impingement wall head at

the center of the combustion chamber and attached to the cylinder head as shown in Fig.

2.8. The effects of combustion chamber geometry on combustion characteristics, engine

performance and exhaust gases are also investigated. The different multii-impingement

wall head and various types of combustion chambers used in the experiments are shown

in Fig. 2.9 and 2.10 respectively. They found that the reentrant type of combustion

chamber with a projection and cutout has a better fuel consumption and lower harmful

emissions. They also found from the photographs that the fuel spray is better diffused and

distributed. This is because the engine can obtain a higher squish in the above case. This

leads to a higher airflow by the micro turbulence in the compression stroke and the back

squish in the power stroke for improved performance.

17

Fig. 2.8 Concept of MIW head for the NICS-MH engine (Lin et al., 1995)

Fig. 2.9 MIW head used in the experiments (Lin et al., 1995)

Fig. 2.10 Four combustion chambers used in the experiments (Lin et al., 1995)

18

c. Fuel injection process and combustion chambers

i) High pressure injection

Corcione at al (1991) in their experiments examined the effects of spray angle,

holes diameter and number, compression ratio and the combustion chamber geometry on

engine performance and emissions. At high speeds sacless nozzles used in reentrant bowl

showed reduction in HC and NOx with unchanged BMEP and BSFC under certain

condition. But at low engine speed torroidal bowl gave better results.

Takeda at al (1996) in their study advanced the fuel injection timing and operated

the engine with the premixed lean Diesel Combustion (PREDIC) to promote fuel air

mixing. They reported that with the PREDIC operation a luminous flame was not

observed during the main combustion period due to improved fuel air mixing. They

concluded that there was a reduction in NOx because the fuel air mixing is made leaner

and the stochiometric ratio mixture in the combustion region is reduced. Also, HC and

CO levels increased because of the fuel air mixture was over lean.

Fig. 2.11 Fuel spray location of MULDIC (Hashizume et al., 1998)

19

ii) Auxiliary Gas / Air Injection Processes

Konno et al (1992) attempted to reduce smoke emitted from direct injection diesel

engine by generating strong turbulence during combustion process. For this purpose a

small auxiliary chamber and fuel injection nozzle were installed at the cylinder head of

the basic engine (refer Fig.2.12) which is termed as the combustion chamber disturbance

(CCD). In CCD a small amount of fuel is injected by using separate injection pump. Four

different diameters (2, 4, 6 and 8mm) of the passage connecting the CCO and main

chamber were investigated. EGR and water injection into the intake manifold were also

examined with the CCO system. They concluded that smoke reduction becomes large

with higher jet momentum and a combination of EGR of water injection with CCD is

very effective to achieve simultaneous reduction of both NOx and also in present system

water is injected at high loads and EGR at low loads.

Fig. 2.12 Cross section of the CCD system (Konno et al., 1992)

Choi et al (1995) investigated the effect of introducing a gas jet. In this case they

tried with industrial nitrogen and carbon dioxide with advanced and retarded timing of the

fuel injection, at a particular timing in the cylinder during the later part of the diesel

combustion. The arrangement of the gaseous injector in the head is as shown in Fig. 2.13.

20

Fig. 2.13 Cylinder head configuration (Choi et al., 1995)

They concluded that the reduction of particulate was controlled by a combination

of the total momentum input and the specific timing at which the momentum was

introduced. For both retarded fuel and gaseous injection timing, higher the jet momentum

the larger is the soot reduction. When injecting CO2 at retarded timing, the rate of

reaction for the carbon-carbon dioxide reaction was too small for any soot oxidation by

CO2 to occur. They also reported that the reduction in NO emissions is caused by ceasing

the NO formation by creating local lower temperature region.

Kurtz et al (2000) used auxiliary gas injection (AGI) to increase in-cylinder mixing

during the latter portion of the combustion in a DI diesel engine in order to reduce soot

emissions without affecting NOx. The equipped auxiliary gas injector for injecting either

nitrogen or air in three different directions 0, 45 and 90 from the center of the combustion

chamber respectively.

d. Bluff bodies

Igarashi (1999) investigated the performance of the vortex shedders as shown in

the Table 2.4. They reported that the vortex shedding caused by the circular cylinder with

a slit and the triangular-semicircular cylinder is excellent in regularity and intensity as

21

compared to that of the ordinary trapezoidal cylinder and concluded that the circular

cylinder having a slit corresponds to d/D=O.2-0.267 and s/d=.l is the most efficient vortex

shedder.

Possibly taking clue from role of vortex shedder in engine, Tanabe et al (2001)

used bluff body as a vortex generator in the combustion chamber of a DI diesel engine

and investigated the engine performance and the exhaust emissions. The also performed a

2-D unsteady computer simulation to classify the effect of the size and shape of the bluff

body and compared with the experimental results obtained in the wind tunnel

experiments. The bluff bodies were set in the piston cavity as shown in Fig. 2.15. They

found that for both bluff body operation unburned emissions CO, THC, NOx and SOF are

lower than non-bluff body operation at low load region.

e. CFD analysis for turbulance

Lisbona et al (2000) studied the process of fuel spray/wall interaction flame

propagation and interaction with the piston surfaces and the most relevant mechanisms of

soot formation and oxidation through CFD analysis to guide the plan of experiments.

They studied two engine operating conditions viz.

a. Maximum power operation and quantified the effect of combustion chamber

geometry on efficiency

b. The emission test cycle.

They analytically proposed a new combustion chamber having a small bowl

which leads to higher swirl levels during expansion accompanied by more soot oxidation

and slightly lower combustion efficiency.

22

2.3 CLOSURE

For improving performance and emission characteristics of a direct injection

diesel engine, the two key aspects identified in this work include using fuel with additive

for better combustion and modifying in-cylinder flow field through turbulence

inducement providing better mixture formation. The fuel additive is expected to alter the

physical and chemical characteristics of the fuel resulting in the reduction of fuel

consumption and/or emissions. The literature on fuel additives reveal that use of aromatic,

metallic and organic additives is widely reported. Many additives serving specific

purposes in the engine on use are found to add to the fuel cost. In certain refinery

processes, the availability of polymer based additive as a bye product could eliminate the

cost consideration in their production. It is also felt that such additives are not thoroughly

investigated.

23

CHAPTER 3

OBJECTIVE OF THE PRESENT WORK

3.1 MOTIVATION

Diesel engines being inferior in their emission characteristics are under greater

scrutiny. Intensive research efforts are ongoing for improving fuel consumption and

exhaust emissions from diesel engines. Due to transient and heterogeneous nature of

diesel combustion, it is imperative to have proper spatial distribution of the injected fuel

and its mixing with air. The role of in-cylinder turbulence in mixture formation and hence

the combustion and emissions is well recognized. Alternatively improvement in the

ignition characteristics of the fuel through additive also yields fuel consumption and

emission advantage. The extensive literature survey suggests rather complex engine

modifications for in-cylinder turbulence inducement and recommends use of metallic and

aromatics based additives. The investigations on turbulence inducement for altering the

fuel-air mixing and polymer base additives for improving combustion are found to be

scanty, hence the present investigation.

3.2 OBJECTIVE

It is observed that investigation on polymer based additive for improving

combustion is not available. However, it is also reported by Flinn (2000) that polymers

can be potential candidates for enhancing performance and emission characteristics.

It is known that enhancing mixing in case of Diesel engines will significantly improve

their performance and reduce emissions. Though several methods have been tried many

of them are quite complex. Thus it is felt that there is a need to develop a method that can

be easily implemented for improving the in cylinder turbulence in a DI Diesel engine.

24

These methods can also be tried in combination with fuel improvements. The present

work addresses these aspects.

Aim of the present work is to investigate the effect of

a. Polymer based additive in diesel and

b. Simple methods to enhance in cylinder turbulence for improving the combustion and

reducing emissions in a DI diesel engine.

25

CHAPTER 4

EXPERIMENTAL WORK

For the present investigation, an experimental set up is installed in the laboratory

with the necessary instrumentation to measure performance, combustion and emissions

from a direct injection compression ignition engine at different operating conditions. A

schematic of the experimental set up is shown in the Fig 4.1. This set up involves an

engine with dynamometer and provisions for measurement of engine speed, fuel and air

flow rates, cylinder pressure history and exhaust emissions such as smoke and oxides of

nitrogen. The details of each of these components of the test set up are furnished below.

In this chapter the experimental set up and its instrumentation are discussed in detail.

4.1 TEST ENGINE

A single cylinder air cooled four stroke, direct injection (DI) compression ignition

diesel engine is chosen for the present investigation. The detailed engine specifications

are provided in Table 4.1. The engine is fitted with conventional fuel injection system,

which has a three orifice of 0.24 mm separated at 120 degrees, inclined at an angle of 60

degrees to the cylinder axis. The recommended injection timing by the manufacturer is 28

deg bTDC (static) 5d the nozzle opening pressure of 190 bar .A centrifugal governor

fitted on the engine enables automatic regulation of the engine. The engine operates at a

constant speed of 1500 rpm. The engine has a hemispherical combustion chamber with

the overhead valve arrangements operated by push rods. The air required for the engine

cooling is forced by the cowl on to the fins, which are present on the periphery of the

cylinder wall. A provision for in cylinder pressure measurement is made on the cylinder

head to mount the piezoelectric transducer.

26

Fig. 4.1 Schematic of Experimental Setup

Table 4.1 Engine specifications

No. of cylinders : vertical

Cylindrical axis : 1

Bore : 0.095 m

Stroke : 0.110 m

Displacement volume : 0.000780 m3

Compression ratio : 15.6:1

Arrangement of valve : overhead

Rated output : 5.5 kW @ 1500 rpm

Speed : 1500 rpm

Cooling system : air-cooled

Fuel Injection timing : 28 deg bTDC

Valve timing

Inlet valve opening : 12° bTDC

Inlet valve closing : 33° aBDC

Exhaust valve opening : 38° bBDC

Exhaust valve closing : 3° aTDC

Valve overlap period : 15° Crank Angle

27

4.2 ENGINE INSTRUMENTATION

The details of the engine instrumentation associated with the present test set up

are discussed below.

4.2.1 Pressure measurement

A piezoelectric transducer is commonly preferred due to its small size, quick

response and accuracy. The transducer used in engine testing needs to have very high

natural frequency for its mechanical vibrations compared to the frequencies of pressure

waves in the engine cylinder and other noise/vibrations in order to avoid resonance and

pickup of other noise. Charge is created on the surface of the transducer when it is

subjected to pressure. Both transverse and longitudinal charge can be created. The

transverse charge created is found to correlate linearly with pressure applied. The charge

produced by the pressure transducer is converted to analog voltage reading by a charge

amplifier.

The pressure transducer has to be calibrated using a dead weight instrument

resulting in a linear relationship between pressure and voltage represented as

Pressure = B * Voltage + C

where B is the slope and termed as calibration constant and C is the intercept of the curve

on the pressure axis.

4.2.2 TDC encoder

An electro optical sensor is fabricated and used to give voltage pulse exactly when

the TDC position is reached. This sensor consists of a well aligned pair of infrared diode

and photo transistor so that infrared rays emitted from the diode fall on the photo

transistor when uninterrupted. A thin metal plate was fixed to the flywheel such that it

passes through the slit between the optical sensor and the infrared diode when engine is

28

running. The vertical plate and the sensor are positioned in such a way that when the

piston reaches TDC, the upper edge of the plate cuts the light emitted by the diode and the

output voltage from the photo-transistor to 5 volts. Voltage signals from the optical sensor

were fed into the analog to digital converter and then data acquisition system along with

the pressure signals for recording.

4.2.3 Analog to Digital converter

Engine Cylinder pressure and TDC signals are acquired and stored on a high

speed computer based digital data acquisition system. A 12 bit analog to digital (A/D)

converter was used to convert analog signals to digital forms. The A to D card had

external and internal trigger facility and with sixteen ended channels. During

experiments, data from 100 consecutive cycles are recorded and signals are then passed

with specially developed software to obtain the combustion parameters and also the heat

release.

4.2.4 Power measurement

The engine is coupled with the swinging field electrical dynamometer. It is

basically a shunt motor that can operate as a generator and a motor. A photo sensor along

with the digital rpm indicator is used to measure the speed of the engine. The voltage

pulses from the sensor are sent to the digital speed meter for pulse conversion and display

of the engine speed with an accuracy of 1 rpm.

4.2.5 Fuel flow rate measurement

Fuel flow rate was measured on the volume basis using a burette and stopwatch.

The fuel from the tank is sent to the engine through a graduated burette using a two way

valve. When the valve is set at position 1 the fuel is sent to the engine directly and in

position 2 the fuel contained in the burette is sent to the engine. For the measurement of

29

the fuel flow rate of the engine, the valve is set at position 2 and the time for a definite

quantity of the fuel flow is noted. This gives the fuel flow rate for the engine.

4.2.6 Air flow rate measurement

The inlet manifold of the engine is connected to the surge tank to avoid pressure

fluctuation at the inlet. A calibrated turbine type flow meter is attached to the tank which

is directed to the atmosphere. This is done with due care that there is no air leakage.

During the engine operation the air to the engine from the atmosphere is through the flow

meter. The time required for the intake of a definite quantity of air gives the airflow rate

of the engine.

4.2.7 Temperature measurement

Temperature of the exhaust and the mean cylinder wall temperature were

measured using chromel-alumel (k-Type) thermocouples. The thermocouple wire

diameter is 2mm and the bead diameter is 5mm. A digital indicator with automatic room

temperature compensation facility was used. For the cylinder wall temperature

measurement the thermocouple was located on the outer surface of the cylinder wall. The

temperature indicator was calibrated periodically.

4.2.8 Smoke measurement

Smoke level was measured using a standard BOSCH smoke meter system. The

measuring instrument consists of a sampling pump that sucks a definite quantity of 330

cm3 of the exhaust sample through a white filter paper. The smoke particle gets deposited

on the filter paper due to which it become coloured. Before every measurement, it was

ensured that the exhaust sample from the previous measurement was completely removed

from the sampling tube and the pump. This sample is then taken to the test bench for

being tested with the BOSCH smoke meter. This consists of the light source and a annular

30

photo detector surrounding it. This instrument sends out a light beam of a calibrated

intensity. The reflected light intensity is determined using a photoelectric cell. This

instrument is calibrated to read zero with a white paper and 10 with a completely black

paper. Before every measurement, the smoke meter is calibrated for zero reading using a

plain white filter paper. The reflectivity of the filter paper gives the smoke value of the

collected sample.

4.2.9 Measurement of oxides of nitrogen

Nitric oxide emission in the exhaust gas is measured with the chemiluminescent

analyzer. This method for detection of NO is based on reaction of NO with ozone to

produce nitrogen dioxide and· oxygen. The N02 molecules from their electronically

excited state revert to the ground state with the emission of photons in the wavelength of

0.6 to 3 µm and are measured by the photomultiplier tube. These photons are directly

proportional to the NO concentration. For the measurement of total oxides of Nitrogen the

N02 is first converted to NO using a converter and the total measure of NO and NO2

together which is termed the total oxides of nitrogen.

NO + O3 NO2 * + O2

NO2 * NO2 + hv

where h is plank’s constant and v is the frequency (Hz).

4.3 ENGINE EXPERIMENTATION

The engine experimentation in this work required two independent sets of

experiments, one involving the modifications in injected fuel using polymer based

additives and another involving the modifications in the engine combustion chamber. In

this two aspects of experimental investigations, the measurements concerning

31

performance, combustion and emissions are carried out on the instrumented test engine

discussed above. The details of these experimental tasks are given below one by one.

4.3.1 Experiments with fuel modifications using additives

The experiments using additives with fuel are carried out at a constant engine speed of

1500 rpm with varying loads between no load to full load conditions. The tests are

performed with six different polymer base additive varieties different physical and

chemical properties. The additive is poly iso butylene (PIB) base polymer and

information like structure of the additive could not be given due to their propriety nature.

The samples of the additives are prepared in a solvent viz. the mineral turpentine oil

(MTO) in the ratio of 30% to 70% by mass respectively. This preparation of the polymer

additive is necessitated to improve their miscibility with the fuel. The mixture of MTO

and the pure additive is termed the additive, which is added to the base diesel fuel in all

the experiments conducted for this purpose. These samples of additives with MTO are

then added to the commercial diesel fuel in proportions of 0.5%, 1.0%, 1.5% and 2.0%

for tests. Thus, 4 proportions of 6 additives, provide 24 test samples for investigation.

4.3.2 Experiments with engine modifications

This arrangement is based on the fact that any disturbance in the flow will

influence the in-cylinder air motion within engine chamber and hence the fuel-air mixing.

In the present investigation, the modifications in the engine are intended to induce

turbulence in the engine combustion space. For this purpose, the use of techniques of

generating turbulence either by inserting bluff bodies or producing jets have been made.

The arrangements investigated in the work are described below.

32

4.3.2.1 Turbulence inducement through bluff bodies

In the present case the arrangement of two cylindrical rods of 3 mm diameter

across bowl in the piston, as shown in Fig. 4.2, is used. These rods are placed parallel or

perpendicular to the piston pin axis. It is felt that the spray impingement on the bluff

bodies is dominant and hence in another arrangement, the orientation of rod is chosen

such that the spray impingement on the rods is avoided. This position is determined

through a bench test of the injector by capturing the spray impressions on the plain paper.

The spray is then superposed with the dimensions of the piston bowl and the piston

diameter and the point of injection. It is found that a single rod fixed at a position of 40

degrees anticlockwise to the piston pin axis will have a minimal spray impingement and

the subsequent experiments are done with this orientation of the rod.

In another arrangement seeking the effect of grooved rods for altering turbulence

level, the plain rods wound with kanthel wire of 0.9 mm thickness are used and the

experiments are done for the three positions that is parallel, perpendicular and at an angle

to the piston pin axis.

Fig. 4.2 Arrangement of bluff bodies on different orientations

33

Table 4.2 Dimensions of the elements used for generating in cylinder turbulence

Case Elements Distance

from

Pin axis

Length

(mm)

Diameter

(mm)

Spacing

(mm)

No.

of

turns

No. of

holes @

angle

1

Parallel

Rod 1

Rod 2

13.5 52.5

46.5

3

2

Perpendicular

Rod 1

Rod 2

18 52.5

52.5

3

3

Angular

Rod 13.5

13.5

58.5 3

4

Parallel with

wire

Rod 1

Rod 2

0.9 2.5 16

14

5

Perpendicular

with wire

Rod 1

Rod 2

0.9 2.5 16

16

6

Angular with

wire

Rod 0.9 2.5 18

7 Internal

jets

3 2@90

8 Internal

jets

3 3@120

9 Internal

jets

3 4@90

10 Internal

jets

3.5 2@90

4.3.2.2 Internal jets

The air present in the combustion chamber can be forced into the bowl space

through the hole drilled from the piston top land, to induce turbulence through the internal

jets so produced. This arrangement should particularly prove useful in enhancing

combustion during the period of fuel injection near the end of compression stroke. The

intensity of turbulence produced by the internal jets so formed would depend on the

position, size and the number of jets in a given arrangement. In order to evaluate the

effects of these parameters, the experiments are conducted on several configurations of

internal jets, given in Table 4.3.

34

Table 4.3 Several configurations of internal jets

Test

cases

Description

1 Two internal jets of 3 mm diameter positioned 180 degrees to each other in cavity

2 Three internal jets of 3 mm diameter positioned 120 degrees to each other in cavity

3 Four internal jets of 3 mm diameter positioned 90 degrees to each other in cavity

4 Two internal jets of 3.5 mm diameter positioned 180 degrees to each other in cavity

35

CHAPTER 5

ANALYSIS PROCEDURE

The following paragraphs describe the procedure adopted for the analysis of the

experimental data obtained during this investigation.

The engine processes, terms and the important parameters necessary for the

performance analysis and their implications are described below.

A schematic representing engine kinematics is given in fig. 5.1 where slider crank

mechanism converts reciprocating motion of piston to the rotary motion of the shaft. The

distance 's' shown in the figure is given by the equation

S = a cos + (l2 – a

2 sin

2)½

where a is crank radius, l is connecting rod length, s is instantaneous piston position and

is instantaneous crank angle.

Fig. 5.1 Slider crank mechanism of an IC engine

36

The thermodynamic cycle of a four-stroke diesel engine consists of four important

processes.

i. Intake (IVO-IVC)

ii. Compression (IVC-SFI-TDC)

iii. Combustion and expansion (SFI-SIGN-ECOMB—EVO)

iv. Exhaust (EVO-EVC)

These events of the engine are represented on a pressure-volume diagram in Fig. 5.2.

Fig. 5.2. Pressure volume diagram of a thermodynamic cycle

As compression starts, both the curves begin close together but the fired engine

pressure starts separating out gradually from the motoring curve on account of

combustion energy release. This drift from motored diagram enables estimate of the

ignition delay which is the period elapsed between the start of injection to the onset of

combustion.

37

The following useful performance characteristics are estimated from the measured values.

Brake thermal efficiency: (BTH) BTH = output power / (FC*CV)

Brake specific fuel consumption: (BSFC) BSFC=FC/BP

Fuel consumption: (FC)

FC= [known quantity of fuel consumed / time taken for the known quantity of the fuel

to be consumed] * density of the fuel

Brake power: (BP) BP=W*N/C

where W is the load on the dynamometer, N is the speed of the engine, C is the

dynamometer constant.

The following two aspects are needed through analysis in order to explain the

experimental results of these investigations.

i. Combustion analysis

ii. Mixing / Turbulence analysis

5.1 COMBUSTION ANALYSIS

The instantaneous experimental data are acquired over several cycles. For averaging,

pressure data of approximately 100 thermodynamic cycles are chosen. The first rise in the

voltage signal due to IDC indicator is taken as a IDC position. At a fixed clock frequency

of the data acquisition card of 100 kHz, approximately 370-380 pressure-voltage readings

are acquired by the PC for each rotation of the crankshaft. By interpolation, the pressure-

voltage readings are arranged at a spacing of 1 CA degree. The interpolation is more

accurate, if done through spline fitting. Since the engine is four-stroke type, 720 such

interpolated data correspond to one complete thermodynamic cycle (intake, compression,

combustion and exhaust) of the engine. The interpolated data are corrected for the

transducer drift by subtracting from them, a linearly increasing voltage (-2mV/s).

Subsequently these data is multiplied by the constant "B" to obtain it into relative

38

pressure values at each instant. These pressure data are required to be referenced using a

particular known pressure, hence pressure at inlet BDC is taken equal to the inlet

manifold pressure.

5.2 MIXING / TURBULENCE ANALYSIS

One of the aspects of investigations carried out in the present work relates to the

in-cylinder turbulence inducement and assessing its consequent effects on engine

performance and combustion· characteristics through the measured pressure-time

diagram. Hence along with combustion analysis procedure described above, a method to

evaluate the changes in the turbulence level affecting fuel-air mixing/combustion

becomes necessary. For the purpose of these interpretations about turbulence parameter

needed in this work, a detailed three dimensional fluid dynamic analysis became

imperative. From the IC engine applications stand point the licensed CFD package-ST

AR"-CI) available in the institute is found appropriate and made use of.

The numerical method for STAR-CD (User Guide. 2001) includes the following steps:

i. Approximation of the unknown flow variables by means of simple functions

ii. Discretisation by substitution of the approximations into the governing flow

equations and subsequent mathematical manipulations

iii. Solution of the algebraic equations.

Prior to the use of STAR-CD solver, the geometry of the object has to be created

and meshed. The closed cycle three-dimensional engine simulation involving

compression and expansion strokes is attempted on two different geometries viz. base

engine combustion chamber and that with modifications for internal jets. The total

combustion space is divided into two regions that is piston bowl and outer annular space.

The bowl region is meshed in GAMBIT while the outer annular space is meshed in

PROST AR and then merged together. Since there are no valves all open surfaces are

39

taken as wall boundary condition with no slip. Initial pressure and temperature inside the

cylinder is assumed as 1 bar and 293 K. The initial velocity is taken as zero. When the

meshed object is imported in the STAR-CD, the mesh can be made to translate, rotate or

distort in any prescribed way, by specifying time- varying positions for some or all of the

cell vertices due to its general dynamic meshing capabilities. Some practical applications

of moving meshes require a large variation in the solution domain size. STAR-CD

overcomes these potential problems by enabling cells to be removed or added during the

transient calculation. Thus, the average cell size can remain roughly constant. The general

approach in cell removal is that mesh motion causes two or more opposing pairs of cell

faces to become coincident at a specified time step, thereby causing all other faces to

collapse to lines or points and thus making the cell disappear. The mass, momentum and

energy associated with collapsed layer will be added to neighboring layer in

volumetrically conservative manner.

The opposite process is used for cell addition, i.e. a previously removed cell

(taken out either during or prior to the fluids calculation) is made to reappear. The initial

conditions for added layer are extrapolated from neighboring layer. This capability of cell

removal and addition was used extensively in this project. The general methodology for

cell activations and deactivations is that of specifying 'events'. Each event is associated

with a unique time step. When the simulation time matches with the event time, that

particular event is executed. The actual mesh movement is specified in terms of the latest

vertex positions in another file called 'cgrid'. In the present work since all cases requires

mesh movement i.e. addition or removal of cell layers, structured hexahedral cells has

been used.

40

CHAPTER 6

RESULTS AND DISCUSSION

The present investigation concerns improvement in performance, combustion and

emission characteristics of a direct injection diesel engine through methods enabling

improvement in ignition characteristics of the fuel by use of additive, and achieving better

fuel air mixing by turbulence inducement. A combined effect of best of both the cases

viz. the polymer base additive and the turbulence inducement through internal jet

arrangement are also evaluated. The experimental results obtained and the combustion

and turbulence analyses carried out during this investigation forms the basis of the

discussions presented in the fol1owing paragraphs.


The tests are performed with six different polymer based additives. The samples

are the mixture of pure additive and solvent i.e. Mineral Turpentine oil (MTO). It is a

form of hydrocarbon (Mineral Turpentine Oil) primarily used to enhance the mixability of

the pure additive with the diesel fuel. However the effect of the pure Mineral Turpentine

oil could not be investigated independently due to its non-availability. In this mixture the

percentage of pure additive is 30% by volume. In the tests conducted with the mixtures

(cal1ed additive in this work) in proportions of 0.5%, 1.0%, 1.5% and 2.0"10 the

proportion of the polymer based additive is 0.15%.0.3%, 0.45% and 0.6% respectively.

The experimental results obtained in this work using different additives (designated

additive 1-6 in Table) mixed with the base diesel fuel in varying proportions of 0.5, 1.0,

1.5, and 2.0 percent are discussed. Figures 6.1- 6.96 show various experimental results

concerning the variations of the engine performance, combustion and emissions with

different additives and their comparison with the base diesel fuel. The respective

41

combinations of different additives at different conditions are indicated in each figure.

The important observations from the results include:

Brake Specific Fuel Consumption (Figs. 6.1 - 6.6) is found to improve in the

case of all additives at almost all loads. At full load, the improved BSFC value ranges

between 250 g/kWh to 268 g/kWh for most of the additives at their various percentage

additions as compared to the base fuel value of 288 g/kWh. The best improvement of 39

g/kWh as against base fuel value of 288 g/kWh. For additive mixture I and 6 at 0.5 %

addition, the measured BSFC values are 266 and 268 g/kWh.

Exhaust Smoke (Figs. 6.7-6.12) levels reduced in all the cases at almost all load

conditions. At full load, the smoke values are in the range 3.4 to 5.1 BSU as compared .

to the base fuel value of 5.7. The smoke reduction is the highest (smoke level of 3.4 aSU)

for 1.0% addition of additive mixture 5. The additive mixture 6 at 0.5 % gives a smoke

number of 3.5, which is close to the minimum while additive mixture 1 at 0.5% addition

the measured smoke value is 3.6 BSU. At full load and various percentages of additive

mixtures, the range of smoke values shown in Table 6.1

Table 6.1 Range of variations for different performance parameters at full load

Additive

mixture

Range of

BSFC

Range of

smoke (BSU)

Range of

IMEP values

(bars)

Range of NO

(ppm)

1 253-266 3.6 - 4.2 7.48-7.70 712-721

2 265-269 4.2 - 4.9 7.43-7.82 712-733

3 256-260 3.7 - 4.6 7.46-7.69 711-722

4 250-276 4.0 - 5.2 7.77-7.39 702-724

5 266-280 3.4 - 5.1 7.60-7.71 708-728

6 249-268 3.5 - 4.6 7.20-7.90 699-732

Diesel fuel 288 5.6 7.20 686

42

200

300

400

500

600

0 20 40 60 80 100

LOAD (%)

BS

FC

(g/k

W h

)

0.50% 1.00%

1.50% 2.00%

BASE

Fig.6.1 Variation of BSFC with load (additive - 1)

200

300

400

500

600

0 20 40 60 80 100

LOAD (%)

BS

FC

(g/k

W h

)

0.50% 1.00%

1.50% 2.00%

BASE


43

200

300

400

500

600

0 20 40 60 80 100

LOAD(%)

BS

FC

(g

/kW

h)

0.50% 1.00%

1.50% 2.00%

BASE


200

300

400

500

600

0 20 40 60 80 100

LOAD (%)

BS

FC

(g

/kW

h)

0.50% 1.00%

1.50% 2.00%

BASE


44

200

300

400

500

600

0 10 20 30 40 50 60 70 80 90 100

LOAD (%)

BS

FC

(g/k

W h

)0.50% 1.00%

1.50% 2.00%

BASE


200

300

400

500

600

0 20 40 60 80 100

LOAD(%)

BS

FC

(g/k

W h

)

0.50% 1.00%

1.50% 2.00%

BASE


45

0

1

2

3

4

5

6

0 20 40 60 80 100 120

LOAD(%)

SM

OK

E N

o (

BS

U)

0.50% 1.00%

1.50% 2.00%

BASE

Fig.6.7 Variation of smoke with load (additive - 1)

0

1

2

3

4

5

6

0 20 40 60 80 100

LOAD(%)

SM

OK

E N

o (

BS

U)

0.50% 1.00%

1.50% 2.00%

BASE


46

0

1

2

3

4

5

6

0 20 40 60 80 100

LOAD (%)

SM

OK

E N

o (

BS

U)

0.50% 1.00%

1.50% 2.00%

BASE


0

1

2

3

4

5

6

0 20 40 60 80 100LOAD (%)

SM

OK

E N

o (

BS

U)

0.50% 1.00%

1.50% 2.00%

BASE


47

0

1

2

3

4

5

6

0 10 20 30 40 50 60 70 80 90 100LOAD(%)

SM

OK

E N

o (

BS

U)

0.50% 1.00%

1.50% 2.00%

BASE


0

1

2

3

4

5

6

0 20 40 60 80 100

LOAD (%)

SM

OK

E N

o (

BS

U)

0.50% 1.00%

1.50% 2.00%

BASE


48

Exhaust NO (Figs. 6.13-6.18) - In general with additive mixture, the

concentration of exhaust NO is found to increase from base fuel values (686 ppm). At full

load, the NO values are in the range 699 to 733 ppm. The lowest NO value is observed at

1 % addition of additive mixture 6 and the highest at 1.5% addition of additive mixture

2.The additive mixture s 6 and I gave exhaust NO concentration of 721 and 712 ppm

respectively at 0.5% addition to the base diesel fuel.

Cylinder gas temperature (Figs 6.19 - 6.24) is increased with all the different

additive mixture in various proportions added to the base diesel fuel.

Indicated mean effective pressure (Figs.6.25-6.30) have increased in all cases at

almost all loads and is the highest for additive mixture 6. The full load IMEP value with

base fuel is found to be 7.20 bars. At full load, the IMEP value of additive mixture 6

ranges between 7.2-7.9 bars at various percentage additions. It shows the highest IMEP of

7.89 bars at 2 % addition. At full load and various. percentages of additive mixture, the

ranges of IMEP values are shown in Table 6.1.

Peak rate of pressure rise (Figs. 6.37 - 6.42) decrease in all cases above 20%

load conditions by almost the same amount compared to the base engine value making

combustion somewhat smoother.

Peak cylinder pressure (Figs. 6.43 - 6.48) decreases at various load conditions in

almost all percentages of additive mixture s by about 1 or 2 bars and the occurrences of

peak pressures delayed by 1 to 2 °CA.

49

100

200

300

400

500

600

700

800

0 20 40 60 80 100

LOAD (%)

NO

(ppm

)

BASE a105

a110 a115

a120

Fig.6.13 Variation of NO with load (additive - 1)

100

200

300

400

500

600

700

800

0 20 40 60 80 100

LOAD(%)

NO

- p

pm

BASE a105

a110 a115

a120


50

100

200

300

400

500

600

700

800

0 20 40 60 80 100LOAD(%)

NO

-pp

m

BASE a105

a110 a115

a120


100

200

300

400

500

600

700

800

0 20 40 60 80 100LOAD(%)

NO

-ppm

BASE a105

a110 a115

a120


51

100

200

300

400

500

600

700

800

0 20 40 60 80 100

LOAD(%)

NO

-pp

m

BASE a105

a110 a115

a120


100

200

300

400

500

600

700

800

0 20 40 60 80 100

LOAD(%)

NO

-pp

m

BASE a105

a110 a115

a120


52

0

500

1000

1500

2000

2500

210 240 270 300 330 360 390 420 450 480 510 540

CRANK ANGLE-deg

CY

LIN

DE

R T

EM

P.

-K .

BASE

a110100

Fig.6.19 Variation of cylinder gas temperature with crank angle (additive - 1)

0

500

1000

1500

2000

2500

210 240 270 300 330 360 390 420 450 480 510 540

CARNK ANGLE-deg

CY

LIN

DE

R T

EM

P.

-K .

BASE

a210100


53

0

500

1000

1500

2000

2500

210 240 270 300 330 360 390 420 450 480 510 540

CRANK ANGLE - deg

CY

LIN

DE

R T

EM

P.

-K .

BASE

a310100


0

500

1000

1500

2000

2500

210 240 270 300 330 360 390 420 450 480 510 540

CRANK ANGLE-deg

CY

LIN

DE

R T

EM

P.

-K .

BASE

a410100


54

0

500

1000

1500

2000

2500

210 240 270 300 330 360 390 420 450 480 510 540

CRANK ANGLE-deg

CY

LIN

DE

R T

EM

P.

-K .

BASE

a510100


0

500

1000

1500

2000

2500

210 240 270 300 330 360 390 420 450 480 510 540

CRANK ANGLE-deg

C

YL

IND

ER

TE

MP

. -K

. BASE

a610100


55

1

2

3

4

5

6

7

8

0 10 20 30 40 50 60 70 80 90 100

LOAD (%)

IME

P(b

ar)

0.005 0.01

0.015 0.02

BASE

Fig.6.25 Variation of IMEP with load (additive - 1)

1

2

3

4

5

6

7

8

0 20 40 60 80 100

LOAD(%)

IME

P(b

ar)

0.50% 1.00%

1.50% 2.00%

BASE


56

1

2

3

4

5

6

7

8

0 20 40 60 80 100

LOAD(%)

IME

P(b

ars

)

0.50% 1.00%

1.50% 2.00%

BASE


1

2

3

4

5

6

7

8

0 20 40 60 80 100

LOAD(%)

IME

P(b

ars

)

0.50% 1.00%

1.50% 2.00%

BASE


57

1

2

3

4

5

6

7

8

0 20 40 60 80 100

LOAD(%)

IME

P(b

ars

)

0.50% 1.00%

1.50% 2.00%

BASE


1

2

3

4

5

6

7

8

0 20 40 60 80 100

LOAD(%)

IME

P(b

ars

)

0.50%1.00%1.50%2.00%BASE


58

Maximum cycle temperature (Figs 6.31 - 6.36) increased for all the additive

mixture with different proportions of additive mixture with the base diesel fuel which

provide the reason for the increase in exhaust NO concentrations.

Combustion duration (Figs. 6.37 -6.42) reduced by about 10 °CA at full load and

by about 5°CA at part load conditions. The lowest combustion duration (CD) is observed

at all percentages of additive mixture I. The decrease is about 13 °CA at full load

condition with additive mixture I. There is, in general, decrease in combustion duration in

the presence of the additive mixture.

Exhaust gas temperatures (EGT) (Figs. 6.43 - 6.48) at different loads are

generally lower than base fuel value except at full load where the exhaust gas

temperatures increased by 10 - 45 0 C compared to the base fuel case. The decrease is

found to be the highest for additive mixture 6 at almost all loads except full load. The

higher is the percentage addition of an additive, the greater is the decrease in EGT.

Cylinder wall temperatures (Figs. 6.49 - 6.54) decreased by about 5-10 °C at all

loads and additive percentages compared to the base engine value. This may be due to

lower heat transfer to the walls and hence lower heat losses.

BSFC (Figs. 6.55 - 6.60) is observed to improve in all the cases and is the lowest

in the case of Additive mixture 6 with 2% addition of the additives.

Bosch smoke number (Figs. 6.61 - 6.66) is seen to reduce in all cases and is the

lowest in additive mixture 5 with 1.0% addition of the additive mixture with the base fuel.

59

500

1000

1500

2000

2500

3000

0 20 40 60 80 100

LOAD(%)

Tm

ax-K

0.005 0.01

0.015 0.02

BASE

Fig.6.31 Variation of peak cylinder temperature with load (additive-1)

1000

1500

2000

2500

3000

0 20 40 60 80 100

LOAD(%)

Tm

ax-K

0.005 0.01

0.015 0.02

BASE


60

500

1000

1500

2000

2500

0 20 40 60 80 100

LOAD(%)

Tm

ax-K

0.005 0.01

0.015 0.02

BASE


500

1000

1500

2000

2500

0 20 40 60 80 100

LOAD(%)

Tm

ax-K

0.005 0.01

0.015 0.02

BASE


61

500

1000

1500

2000

2500

0 20 40 60 80 100

LOAD(%)

Tm

ax-K

0.005 0.01

0.015 0.02

BASE


500

1000

1500

2000

2500

0 20 40 60 80 100

LOAD(%)

Tm

ax-K

0.005 0.01

0.015 0.02

BASE


62

50

55

60

65

70

75

80

85

90

0 20 40 60 80 100

LOAD(%)

Tcd(d

eg C

A)

0.005 0.01

0.015 0.02

BASE

Fig.6.37 Variation of total combustion duration with load (additive-1)

50

55

60

65

70

75

80

85

90

0 20 40 60 80 100

LOAD(%)

Tcd(d

eg C

A)

0.0050.010.0150.02BASE


63

50

55

60

65

70

75

80

85

90

0 20 40 60 80 100

LOAD(%)

Tcd

(de

g C

A)

0.005 0.01

0.015 0.02

BASE


50

55

60

65

70

75

80

85

90

0 20 40 60 80 100

LOAD(% (deg CA)

Tcd

(de

g C

A)

0.005 0.01

0.015 0.02

BASE


64

50

55

60

65

70

75

80

85

90

0 20 40 60 80 100

LOAD(%)

Tcd (

deg C

A)

0.005 0.01

0.015 0.02

BASE


55

60

65

70

75

80

85

90

0 20 40 60 80 100

LOAD(%)

Tcd (

deg C

A)

0.005 0.01

0.015 0.02

BASE


65

0

100

200

300

400

500

600

700

0 20 40 60 80 100

LOAD(%)

EX

HA

US

T G

AS

TE

MP

(d

eg

C) 0.005 0.01

0.015 0.02

BASE

Fig.6.43 Variation of exhaust gas temperature with load (additive-1)

0

100

200

300

400

500

600

700

0 10 20 30 40 50 60 70 80 90 100

LOAD(%)

EX

HA

US

T G

AS

TE

MP

(deg C

) 0.005 0.01

0.015 0.02

BASE


66

0

100

200

300

400

500

600

700

0 20 40 60 80 100

LOAD (%)

EX

HA

US

T G

AS

TE

MP

(deg C

)

0.0050.010.0150.02BASE


0

100

200

300

400

500

600

700

0 20 40 60 80 100

LOAD (%)

EX

HA

US

T G

AS

TE

MP

(d

eg

C)

0.005 0.01

0.015 0.02

BASE


67

0

100

200

300

400

500

600

700

0 20 40 60 80 100

LOAD (%)

EX

HA

US

T G

AS

TE

MP

(d

eg

C) 0.005 0.01

0.015 0.02

BASE


0

100

200

300

400

500

600

700

0 20 40 60 80 100

LOAD(%)

EX

HA

US

T G

AS

TE

MP

(deg C

)

0.0050.010.0150.02BASE


68

60

80

100

120

140

160

180

200

0 20 40 60 80 100

LOAD(%)

CY

LIN

DE

R W

AL

L T

EM

P.

(de

g C

) .

0.005 0.01

0.015 0.02

BASE

Fig.6.49 Variation of cylinder wall temperature with load (additive-1)

60

80

100

120

140

160

180

200

0 20 40 60 80 100

L0AD(%)

CY

LIN

DE

R W

AL

L T

EM

P.

(de

g C

) . 0.005 0.01

0.015 0.02

BASE


69

60

80

100

120

140

160

180

200

0 20 40 60 80 100

LOAD(%)

CY

LIN

DE

R W

AL

L T

EM

P.

(de

g C

) .

0.005 0.01

0.015 0.02

BASE


60

80

100

120

140

160

180

200

0 20 40 60 80 100

LOAD(%)

CY

LIN

DE

R W

AL

L T

EM

P.

(de

g C

) .

0.005 0.01

0.015 0.02

BASE


70

60

80

100

120

140

160

180

200

0 20 40 60 80 100

LOAD(%)

CY

LIN

DE

R W

AL

L T

EM

P.

(de

g C

) .

0.005 0.01

0.015 0.02

BASE


60

80

100

120

140

160

180

200

0 20 40 60 80 100

LOAD(%)

CY

LIN

DE

R W

AL

L T

EM

P.

(de

g C

) .

0.005 0.01

0.015 0.02

BASE


71

200

300

400

500

600

700

0.5 1 1.5 2

% ADDITIVE

BS

FC

(g

/kW

H)

0.2 0.4 0.6

0.8 1

Fig.6.55 Variation of BSFC with percentage additive (additive-1)

200

300

400

500

600

700

800

0.5 1 1.5 2

% ADDITIVE

BS

FC

(g

/kW

h)

0.2 0.4 0.6

0.8 1


72

200

300

400

500

600

700

800

0.5 1 1.5 2

% ADDITIVE

BS

FC

(g/k

W h

)

0.2 0.4 0.6

0.8 1


200

300

400

500

600

700

800

0.5 1 1.5 2

% ADDITIVE

BS

FC

(g

/kW

h)

0.2 0.4 0.6

0.8 1


73

200

300

400

500

600

700

800

0.5 1 1.5 2

% ADDITIVE

BS

FC

(g/k

W h

)

0.2 0.4 0.6

0.8 1


200

300

400

500

600

700

800

0.5 1 1.5 2% ADDITIVE

BS

FC

(g

/kW

h)

20.00% 40.00% 60.00%

80.00% 100%


74

0

1

2

3

4

5

6

0.5 0.7 0.9 1.1 1.3 1.5 1.7 1.9

% ADDITIVE

SM

OK

E B

SU

0.00% 20.00% 40.00%

60.00% 80.00% 100%

Fig.6.61 Variation of smoke with percentage additive (additive-1)

0

1

2

3

4

5

6

7

0.5 1 1.5 2

% ADDITIVE

SM

OK

E B

SU

0 0.2 0.4

0.6 0.8 1


75

0

1

2

3

4

5

6

0.5 1 1.5 2

% ADDITIVE

SM

OK

E B

SU

0 0.2 0.4 0.6

0.8 1


0

1

2

3

4

5

6

7

0.5 1 1.5 2

% ADDITIVE

SM

OK

E B

SU

0 0.2 0.4

0.6 0.8 1


76

0

1

2

3

4

5

6

7

0.5 1 1.5 2

% ADDITIVE

SM

OK

E B

SU

0 0.2 0.4 0.6

0.8 1


0

1

2

3

4

5

6

7

0.5 1 1.5 2

% ADDITIVE

SM

OK

E B

SU

0 0.2 0.4

0.6 0.8 1


77

In order to ascertain a most economical and smoke reducing proportion of the

additive among each of the additive mixture investigated the values of BSFC and exhaust

Bosch smoke number are compared by plotting their values with respect to percentage

additive mixture.

From these experimental results a lower BSFC, reduced smoke and improved and

smoother combustion are observed. A lower BSFC values corroborate with larger fraction

of heat release around TDC with the decreased combustion duration and higher mass

burnt fraction. These experimental results, in turn, suggest a better combustion.

A decrease in peak rate of cylinder pressure and its second rate of change suggest

that the combustion is smoother with additives mixture. A decrease in peak heat release

during the premixed combustion and its occurrence closer to TDC explain the lower

values of peak cylinder pressure, peak gas temperatures and the lower compression work.

It is observed that there is a decrease in peak pressure in almost all cases. However, a

slight increase in calculated peak temperatures could be attributed to an increase in mass

burnt fraction under diffusion combustion in the period close to TDC.

These conditions lead to a better oxidation of fuel, reduced smoke and lower

BSFC. Evidently, the exhaust smoke and BSFC values show similar trends for different

additive mixture at various loads. An increase in diffusion combustion and occurrence of

most of the heat release near TDC tend to provide a greater conversion of heat to work

from the piston, consequently, resulting in a decrease in EGT as observed at almost all

loads (except near full load conditions). This fact corroborates with the observation of

higher mass burnt fraction in the case of most of the additive mixture.

78

Increase in ignition delay can also be attributed to the decrease in exhaust gas

temperatures (EGT) and cylinder wall temperatures (CWT). The lower EGTs and CWTs

would influence the fuel evaporation rate in the initial stages of fuel injection causing

lower peak premixed heat release.

Although in most of the cases the peak cylinder pressures are observed to be

lower, the IMEP values are higher primarily due to the fact that the cylinder pressures

during diffusion phase of combustion in the expansion stroke have maintained higher

values compared to base engine pressures. This again is a consequence of better and short

duration combustion. This vary fact also explains an increase in NO concentration.

The extensive experiments carried out here enable arriving at the effective and

economical proportions of the additive mixture giving a better performance and emission

characteristics. Figure 6.67 shows the variation of the BSFC with load. It can be seen that

in general there is an improvement of 7.6 % is observed in case of 0.5% of additive

mixture 1 in BSFC over the base diesel BSFC value. This is attributed to the better

combustion in the presence of the polymer additive mixture. Figs. 6.68 and 6.69 show the

variation of the peak pressure and the maximum rate of pressure rise with load

respectively in case of 0.5% of additive mixture with samples 1 and 6. The decrease in

peak pressure and the maximum rate of pressure rise in load ranges of 20% to 100%

indicates the smoothness of the engine combustion in the presence of the additive mixture

with diesel fuel. This is due to the fact that in the presence of the polymer additive the

fuel droplets are dispersed more uniformly and the combustion starts much earlier and

bum steadily. Fig. 6.70 shows the variation of the ignition delay at different loads. At part

load conditions there is a decrease in the ignition delay whereas if the engine tends

towards full load operation the ignition delay increases within 1 to 2 degrees crank angle

79

because when the engine tends towards full load more fuel is injected and the prevailing

temperature. In the combustion chamber is not sufficient to start the combustion of the

fuel initially in the presence of the polymer additive mixture which has the tendency to

reduce the overall temperature. Fig. 6.71 shows the variation of the combustion duration

with load and its decrease indicates that the fuel is burnt effectively in a shorter duration

in presence of the additive mixture. This may be due to the formation of very fine

droplets of fuel in the presence of additive mixture. Fig. 6.72 shows the variation of the

energy release rate with load. It can be seen in the graph that there is a delayed occurrence

of the peak energy release, which has shifted slightly towards the TDC position; as a

result there is a decrease in the peak pressure. There is also a decrease in the peak energy

release rate and the diffusion phase of the combustion increased. Fig. 6.71 shows the

variation of the Bosch smoke number with load showing a drastic reduction in the smoke

level of about 38 % in case of the additive mixture 6 (0.5%). This is because of the higher

diffusion phase of the combustion and the shortening of the combustion period, which

ultimately resulted in the reduction of the exhaust smoke. Fig. 6.72 shows the variation of

the NO with load. As a trade-off between oxides of nitrogen and smoke there is a slight

increase in NO emission which is not very significant but very marginal, that is, about 4%

in case of additive mixture 6 at 0.5% addition as compared with the level of reduction in

smoke level at full load over base engine.

80

200

250

300

350

400

450

500

550

600

20 40 60 80 100

LOAD(%)

BS

FC

(g/k

W h

)

BASE

ADDITIVE 1(0.5%)

ADDITIVE 6(0.5%)

Fig.6.67 Variation of brake specific fuel consumption with load

48

52

56

60

64

68

0 20 40 60 80 100

LOAD(%)

PE

AK

PR

ES

SS

UR

E (

ba

r) .

BASE

ADDITIVE 1(0.5%)

ADDITIVE 6(0.5%)

Fig.6.68 Variation of peak pressure with load

81

2

3

4

5

6

7

8

0 20 40 60 80 100

LOAD(%)

MA

X R

AT

E O

F P

RE

SS

UR

E R

ISE

(ba

r/C

A)

BASE

ADDITIVE 1(0.5%)

ADDITIVE 6(0.5%)

Fig.6.69 Variation of maximum rate of pressure rise with load

17

18

19

20

21

22

23

0 10 20 30 40 50 60 70 80 90

LOAD(%)

IGN

ITIO

N D

ELA

Y (

deg C

A)

.

BASE

ADDITIVE 1(0.5%)

ADDITIVE 6(0.5%)

Fig.6.70 Variation of ignition delay with load

82

0

2

4

6

0 20 40 60 80 100LOAD(%)

BO

SC

H S

MO

KE

No

(B

SU

) . BASE

ADDITIVE 1(0.5%)

ADDITIVE 6(0.5%)

Fig.6.71 Variation of smoke number with load

0

200

400

600

800

0 20 40 60 80 100

LOAD(%)

NIT

RIC

OX

IDE

(pp

m)

BASE

ADDITIVE 1(0.5%)

ADDITIVE 6(0.5%)

Fig.6.72 Variation of nitric oxide with load

83

6.2 IN-CYLINDER TURBULENCE INDUCEMENT

In the present investigation, the experimental results are obtained for various

arrangements arising out of the proposed three methods viz. rods, rods with wire and two

internal jets used for inducing turbulence in combustion chamber (refer Table 6.2). The

results concerning the performance, emission and combustion aspects of the engine in

different arrangements of bluff body methods are first compared with the experimental

data of the base engine test. The results of the best of the two bluff body methods, i.e.

rods and rods with wire, are then compared with the experimental results of the base

engine, and the two internal jets in cavity configuration in order to evaluate their relative

effects on the ,engine performance and emissions.

Table 6.2 Various configurations for in-cylinder turbulence inducement

Test cases Description

1 Horizontal rods parallel, perpendicular and inclined at40°

anticlockwise to piston in axis

2 Wire wound horizontal rods parallel, perpendicular and inclined at40°

anticlockwise to piston in axis

3 Two, Three and Four internal jets of 3 mm diameter positioned 180,

120 and 90 respectively degrees to each other in cavity

4 Two internal jets of 3.5 mm diameter positioned 180 degrees to each

other in cavity

6.2.1 Effect of bluff bodies

In the early part of combustion, the energy release rates are observed to be higher

in the different arrangements of simple rods in comparison to that of rods with wire.

Further the arrangement of a single rod on angular orientation gives a higher energy

release rates than the arrangements involving two rods.

84

The presence of bluff bodies alters the clearance volume and heat transfer areas of

the existing engine. In the present case, the change in clearance volume is such that the

maximum change in compression ratio in case of two horizontal rods is about 0.6 percent.

This increase in compression ratio is expected to give an increased cylinder pressures

during compression. On the contrary, the experimental pressure diagrams in the presence

of rods show a reduction in pressure during compression, which at compression TDC is

typically of the order of 4.5 bar. This clearly suggests that there is a dominant influence of

heat loss in the presence of bluff bodies.

In general, these results show an increase in BSFC and smoke values at higher

loads (above 70% load) and a marginal reduction in NOx concentration compared to the

base engine case. The increase in BSFC and decrease in NOx concentrations also

corroborate with the reasoning of increased heat loss in the presence of bluff bodies.

However, the simple rods placed at an angle to the piston pin axis show a decrease in

smoke values at all load conditions. This is possibly an effect of relatively lower heat

losses due to the presence of a single rod and that too at an orientation where spray

impingement is expected to be minimized.

6.2.2 Effect of Internal jets

In the experiments with bluff bodies, the maximum increase in brake specific fuel

consumption is limited to 1.6% of the base engine value. Considering the fact that the

improvements in engine performance due to horizontal bluff bodies are rather low, the

option of inducing turbulence through internal jets has been investigated. In this case, the

two holes are drilled from the flat surface of the piston crown to the cavity such that the

jets of the cylinder charge enter near the bottom of the cavity to facilitate improvement in

mixing. A typical comparison of pressure time histories and energy release rates of two

85

internal jets arrangement with the base engine data and the better of the different bluff

body arrangements with and without wires are shown in Figs. 6.77 and 6.78 respectively.

From Fig. 6.77, it can be observed that the pressure levels of two holes internal

jets arrangement are generally higher to the bluff body cases but the pressure levels still

remain somewhat lower to the base engine case. A comparison of these cases with the

base engine shows a smoother combustion and higher rates of energy release in the later

period of combustion.

The presence of hot internal jets of cylinder charge eliminates the disadvantage of

heat loss otherwise arising in the presence of bluff bodies. The observations concerning

higher values of ignition delays (Fig. 6.73) and the lower peak pressures (Fig. 6. 114)

with a delayed occurrence have remained similar to those observed in the case of bluff

body arrangements discussed earlier (refer Fig. 6.75). These results indicate a superiority

of the jet turbulence over the bluff body turbulence with regard to the engine performance

and exhaust smoke level.

6.3 PARAMETRIC STUDIES

Since the results of the investigations with two internal jets discussed above resulted in

encouraging trends of improvement in engine performance and emission characteristics, it

is felt that some variations in respect of the number, position and size of internal jets are

tested. Among these three parameters the variations of number and position of internal

jets remain coupled. Therefore, the following two sets of experiments are conducted to

examine the effects of

i. number and position of internal jets

ii. size of the internal jets

86

The results based on the limited experiments carried out for the purpose are discussed in

the following section.

16

18

20

22

24

26

0 20 40 60 80 100

LOAD(%)

IGN

ITIO

N D

EL

AY

(d

eg

CA

) . 2 HOLES PARALLEL

PARALLEL WIRE BASE ENGINE

Fig. 6.73 Ignition delay for horizontal rods with and without wire

45

50

55

60

65

70

0 20 40 60 80 100

LOAD(%)

PE

AX

PR

ES

SU

RE

(b

ar)

.

2 HOLES PARALLEL

PARALLEL WIRE BASE ENGINE

Fig. 6.74 Peak pressure for horizontal rods with and without wire, two internal jet along

with base engine

87

200

400

600

0 20 40 60 80 100

LOAD(5)

BS

FC

(g/k

W h

)

2 HOLES

PARALLEL

PARALLEL WIRE

BASE ENGINE

(a)

0

2

4

6

8

0 20 40 60 80 100

LOAD(%)

SM

OK

E N

o (

BS

U)

.

2 HOLES

PARALLEL

PARALLEL WIRE

BASE ENGINE

(b)

0

200

400

600

800

0 20 40 60 80 100

LOAD(%)

NO

x (

ppm

)

2 HOLES

PARALLEL

PARALLEL WIRE

BASE ENGINE

(c)

Fig. 6.75 BSFC, smoke level and NOx for horizontal rods with and without wire, two

internal jet along with base engine

88

6.3.1 Effect of number and position of the internal jets

In the present discussion, the experimental results of modified combustion

chambers having the two, three and four holes internal jets are compared with that of the

base engine performance, combustion and emission characteristics.

Figure 6.76 shows the variation of the brake thermal efficiency with load for two,

three and four internal jets respectively. These results show that there is an increase in

brake efficiency for all the three cases. However, two hole internal jets show the highest

increase of 2% possibly due to better resultant fuel-air mixing in this case compared to

the other two cases. Though there is an increase in brake thermal efficiency in the three

and four holes case compared with base engine, it is found to be somewhat lower

compared to the two holes. This effect is attributed to a possible unfavorable interaction

between the three and four jets due to their closer spacing.

For different number of internal jets, Figure 6.77 shows the variation of the peak

pressure with load. There is a decrease in peak pressures in all these cases compared to

the base engine value with very limited difference in the magnitude of the peak pressures

with different internal jets. Figure 6.78 shows the variation of the maximum rate of

pressure rise with load. The maximum rate of pressure rise is significantly reduced in the

presence of internal jets, suggesting smoother operation of the engine in comparison to

the base engine. Figure 6.79 shows the variation of the ignition delay which is found to

increase in all the cases. These facts relate with each other to corroborate the reduction in

peak pressure and the possible change in mixing level during that phase of combustion.

Figure 6.80 shows the variation of the combustion duration with respect to load. The

combustion duration has become shorter possibly due to better combustion. From Figure

6.87 showing the variation of the energy release rate with load, it can be observed that the

energy release has shifted closer to TDC position and the period of diffusion phase of

89

combustion is greater to be attributed to an increase in combustion rate. There is a

significant reduction in smoke level with the internal jets vis-à-vis base engine smoke

values as shown in Fig. 6.81. The increase in thermal efficiency and shortening of the

total combustion period observed in the case of internal jets are possibly the cause of

decrease in smoke and a marginal increase at full load condition in the concentration of

NO (refer Figure 6.82). Thus, it can be inferred that internal jets could produce more

favorable conditions of in-cylinder turbulence than the basic engine configuration. In case

of three and four jets, there may be interference of internal jets to result in unfavorable

mixing due to level of turbulence produced.

6.3.2 Effect of the size of the internal jets

Figure 6.83 shows that the brake thermal efficiency at various loads is found to

decrease with the increase in jet sizes. However, the increase is more significant with the

3 mm diameter jets than 3.5 mm jets, which is attributed to a better combustion possibly

due to the higher turbulence inducement. The variation of the peak pressure with load for

the different jet sizes along with base engine value is shown in Figure 6.84. In both the

cases of the jet sizes there is decrease in the peak pressure compared with that of the base

engine. The maximum rate of pressure rise is lowered with all the jet sizes as seen from

Figure 6.85, which also suggests that the engine is smoother in operation in the presence

of the internal jets. Figure 6.86 shows the variation of the ignition delay with load and is

found to increase in both the jet sizes. Figure 6.87 shows the variation of the combustion

duration with load. The jets of 3 mm diameter showed shorter combustion duration and

resulted in the better combustion. From the variation of energy release rate with load

(refer Figure 6.92), it is observed that in cases of jets with different sizes there is higher

diffusion phase and the peak shifts towards the TDC. In the case of jets with 3 mm, a

better combustion is observed. Figure 6.88 shows the variation of the smoke level with

90

load. In case of both the jets of diameters 3 and 3.5 mm, there is drastic decrease in

smoke level. The smoke value is lowest with the jets with 3 mm because of the better and

favorable turbulence generation. Figure 6.89 shows the variation of the NO concentration

with load. There is a marginal increase in the NO concentration at full load as a trade-off

effect with the decrease in smoke.

In the CFD analysis of the results of base engine and the case of two internal jets

of 2.5, 3 and 3.5 mm diameter are compared. The parameters like turbulent kinetic

energy, eddy dissipation rate, velocity magnitude and swirl velocity are chosen for

inference on mixing quality. The computed results of these quantities have been plotted in

three crank positions of 30° bTDC, 0° TDC and 30° aTDC respectively.

Figure 6.126 shows the comparison of respective velocity contours indicating an

increase in the swirl velocity component in the case of two internal jets of different

diameters compared to the base engine values. From CFD results, it is also observed that

velocities are higher at the point where internal jets enter the combustion chamber thus

resulting in higher mixing rates due to better distribution of the air pockets in the

combustion chamber.

In summary, Figures 6:135 and 6.136 show instantaneous variations of the mass

averaged turbulent kinetic energy and the eddy dissipation rates respectively from where

it can be clearly seen that due to the presence of jets there is a significant increase in the

turbulent mixing throughout the injection period to provide better combustion and

emission characteristics. This briefly substantiates the useful outcome on engine

performance, combustion and emissions of the internal jet configurations investigated in

the present work. However, this analytical aspect could be a matter of much detailed

investigation in itself. The same was not within the scope of the experimental work

carried out in this investigation.

91

0

10

20

30

40

0 20 40 60 80 100

LOAD(%)

BR

AK

E T

HE

R. E

FF

ICIE

NC

Y(%

) .

TWO JETS

THREE JETS

FOUR JETS

BASE

Fig. 6.76 Variation of brake thermal efficiency with load

45

50

55

60

65

0 20 40 60 80 100

LOAD(%)

PE

AK

PR

ES

SU

RE

(ba

r)

TWO JETS

THREE JETS

FOUR JETS

BASE

Fig. 6.77 Variation of peak pressure with load

92

3

3.5

4

4.5

5

5.5

6

6.5

7

0 20 40 60 80 100

LOAD(%)

MA

X R

AT

E O

F P

R. (b

ar/

de

g C

A)

. TWO JETS

THREE JETS

FOUR JETS

BASE

Fig. 6.78 Variation of maximum rate of pressure rise with load

16

18

20

22

24

40 50 60 70 80 90 100

LOAD(%)

IGN

ITIO

N D

ELA

Y(d

eg C

A)

. TWO JETS

THREE JETS

FOUR JETS

BASE

Fig. 6.79 Variation of ignition delay with load

93

60

65

70

75

80

85

90

0 20 40 60 80 100

LOAD(%)

TO

TA

L C

OM

BU

ST

ION

DU

RA

TIO

N

(deg C

A)

TWO JETS

THREE JETS

FOUR JETS

BASE

Fig. 6.80 Variation of combustion duration with load

0

2

4

6

0 20 40 60 80 100

LOAD(%)

SM

OK

E N

o (

BS

U)

.

TWO JETS

THREE JETS

FOUR JETS

BASE

Fig. 6.81 Variation of smoke number with load

94

0

200

400

600

800

0 20 40 60 80 100

LOAD(%)

NIT

RIC

OX

IDE

(pp

m)

.

TWO JETS

THREE JETS

FOUR JETS

BASE

Fig. 6.82 Variation of nitric oxide with load

0

10

20

30

40

0 20 40 60 80 100

LOAD(%)

BR

AK

E T

HE

R.

EF

FIC

IEN

CY

(%)

.

TWO JETS 3mm

TWO JETS 3.5mm

BASE

Fig. 6.83 Variation of brake thermal efficiency with load

95

35

45

55

65

75

0 20 40 60 80 100

LOAD(%)

PE

AK

PR

ES

SU

RE

(bar)

.

TWO JETS 3mm

TWO JETS 3.5mm

BASE

Fig. 6.84 Variation of peak pressure with load

2

2.5

3

3.5

4

4.5

5

5.5

6

6.5

7

0 20 40 60 80 100

LOAD(%)

MA

X R

AT

E O

F P

RE

SS

UR

E

RIS

E(d

eg/C

A)

TWO JETS 3mm

TWO JETS 3.5mm

BASE

Fig. 6.85 Variation of maximum rate of pressure rise with load

96

14

16

18

20

22

40 50 60 70 80 90 100

LOAD(%)

IGN

ITIO

N D

EL

AY

(de

g C

A)

.

TWO JETS 3mm

TWO JETS 3.5mm

BASE

Fig. 6.86 Variation of ignition delay with load

50

55

60

65

70

75

80

85

90

0 20 40 60 80 100

LOAD(%)

CO

MB

US

TIO

N D

UR

AT

ION

(de

g C

A)

TWO JETS 3mm

TWO JETS 3.5mm

BASE

Fig. 6.87 Variation of combustion duration with load

97

0

2

4

6

0 20 40 60 80 100LOAD(%)

SM

OK

E N

o (

BS

U)

.

TWO JETS 3mm

TWO JETS 3.5mm

BASE

Fig. 6.88 Variation of smoke number with load

0

200

400

600

800

0 10 20 30 40 50 60 70 80 90 100

LOAD(%)

NIT

RIC

OX

IDE

(ppm

) .

TWO JETS 3mm

TWO JETS 3.5mm

BASE

Fig. 6.89 Variation of nitric oxide with load

98

6.4 COMBINED EFFECTS OF FUEL ADDITIVE AND IN-CYLINDER

TURBULENCE MODIFICATIONS

From the investigations carried out in the present work on the fuel additive and in-

cylinder turbulence inducement, the following cases yielded the best of the improvements

in the diesel engine performance, combustion and emission characteristics in the

respective categories:

i. the additive 1 with 0.5 % by volume in the base fuel, and

ii. the two internal jets of 3 mm diameter in the base engine

On realizing the independent effects of the two central aspects investigated in this

work, it is considered necessary that a typical evaluation of the combined effects of these

two independent cases are examined and the net improvements in the engine performance

ascertained. For this purpose, a selective set of experimentation was conducted involving

the aforesaid modifications of fuel with additive and the internal jets combined together

and compared with base engine and their independent test results. The results of these

experiments on BSFC, combustion parameter, exhaust smoke and NO emissions with

varying loads are shown in Figures 6.90 - 6.144.

Figure 6.90 shows a comparison of brake specific fuel consumption (BSFC)

values at various loads obtained for base engine, independent and combined changes of

fuel and fuel-air mixing as identified. It is observed that while fuel and fuel-air mixing

modifications independently yielded 7.6 and 7.8 percent reduction in BSFC, their

combined effects provided 9.8% improvement over the base engine value. This effect is

attributed to the simultaneous effects of additive in fuel which acts as a combustion

enhancer and better fuel dispersion in conjunction with better fuel-air mixing due to

inducement of turbulence in presence of the internal jets.

99

200

250

300

350

400

450

500

550

600

20 30 40 50 60 70 80 90 100LOAD(%)

BS

FC

(g

/kW

h)

BASE

ADDITIVE 1(.5%)

INTERNAL JET (3mm)

COMBINED

Fig. 6.90 Comparison of the brake specific fuel consumption of the best of additive,

internal jet and the combined case with the base engine

40

50

60

70

0 20 40 60 80 100

LOAD(%)

PE

AK

PR

ES

SU

RE

(bar)

.

BASE

ADDITIVE 1(.5%)

INTERNAL JET (3mm)

COMBINED

Fig. 6.91 Comparison of the peak pressure of the best of additive internal jets and the

combined case with the base engine

100

3

3.5

4

4.5

5

5.5

6

6.5

7

0 20 40 60 80 100

LOAD(%)

MR

PR

(ba

r/d

eg

CA

) .

BASE

ADDITIVE 1(.5%)

INTERNAL JET (3mm)

COMBINED

Fig. 6.92 Comparison of the max rate of pressure rise of the best of additive internal jets

and the combined case with the base engine

17

18

19

20

21

22

0 10 20 30 40 50 60 70 80 90 100LOAD(%)

IGN

ITIO

N D

ELA

Y (

deg C

A)

.

BASE

ADDITIVE 1(.5%)

INTERNAL JET (3mm)

COMBINED

Fig. 6.93 Comparison of the ignition delay of the best of additive internal jets and the


101

30

40

50

60

70

0 10 20 30 40 50 60 70 80 90 100

LOAD(%)

CO

MB

US

TIO

N D

UR

AT

ION

(deg C

A)

BASE

ADDITIVE 1(.5%)

INTERNAL JET (3mm)

COMBINED

Fig. 6.94 Comparison of the combustion duration of the best of additive internal jets and

the combined case with the base engine

0

2

4

6

0 10 20 30 40 50 60 70 80 90 100

LOAD(%)

SM

OK

E N

o (

BS

U)

.

BASE

ADDITIVE 1(.5%)

INTERNAL JET (3mm)

COMBINED

Fig. 6.95 Comparison of the smoke number of the best of additive internal jets and the


102

0

200

400

600

800

0 10 20 30 40 50 60 70 80 90 100

LOAD(%)

NIT

RIC

OX

IDE

(ppm

) .

BASE

ADDITIVE 1(.5%)

INTERNAL JET (3mm)

COMBINED

Fig. 6.96 Comparison of the Nitric oxide of the best of additive, internal jets and the


Figures 6.91 and 6.92 show a comparison of the peak pressures and the maximum

rate of pressure rise of the four cases referred here. It is observed that under the combined

effects of modifications in fuel and fuel-air mixing, the peak pressure and the maximum

rate of pressure rise values decreased over base engine values and that obtained during the

independent modifications in fuel and the engine turbulence. These results reveal a

smoother combustion is attained in the engine with the modifications in place. The

variation of ignition delay and combustion durations are shown in figures 6.93 and 6.94

respectively. Generally, a marginal increase of 1° CA is observed in the ignition delay

values over base engine conditions. While the effect on the ignition delay is not

considered too significant, the combustion duration under the combined effect of fuel and

fuel-air mixing modifications seems to have shortened considerably. This would imply

that a better mixture formation due to well timed jet induced turbulence in conjunction

with better combustion with fuel additives have surfaced together for the efficient burning

103

of the fuel and more importantly during the diffusion phase of the diesel engine

combustion.

It appears that the changes in engine combustion have resulted in significant

influence on the exhaust smoke values, as shown in Figure 6.96. The exhaust smoke

values under the combined changes of fuel and fuel-air mixing decreased by 38.5% (5.7

BSU to 3.5 BSU) over the base value. However, the effects of changes in fuel and fuel-air

mixing independently observed to be 36 percent (5.7 BSU to 3.6 BSU) and 20 percent

(5.7 BSU to 4.6 BSU) respectively. As regards nitric oxide emissions (refer Fig. 6.144)

are concerned, a marginal increase of about 5% is observed in case of combined changes

in fuel and fuel-air mixing conditions over the base engine value of NO concentrations at

full load. At part load condition the quantity of the injected fuel is less, which indicates

that a lean combustion occurs at part load condition which resulted in higher NO

emissions whereas at full load the fuel quantity injected is more for the same quantity of

air where the air fuel mixture tends towards stochiometric condition which has resulted in

less NO at full load condition. In conclusion, this selective experimentation has been very

revealing of the combined effects of the two aspects investigated in the present work.

104

Specification of equipments used in engine measurements

Specifications of the pressure transducer and charge amplifier

Pressure Pick up

Make : AVL GRAZ, Austria

Type : 120 QP 250 C, Quartz pressure transducer

Measuring range : 0- 120 bars

Sensitivity : 69.79 pC/bar

Linearity : c+0.5% '

Natural frequency : 67 kHz

Charge Amplifier

Make : AVL GRAZ, Austria

Type : 3056-A01

Output voltage : 0 ≥ ± 10 v at load ≥ 1 k Ω Output current : ± 10 mA

Output impedance : ≤ 0.01 Ω

Digital data acquisition system

Analog to digital converter

Make : DYNALOG-MICROSYSTEMS PVT LTD.,

Type : PCL 818 HG

Number of channels : 16

Internal clock : 1 MHZ

Maximum sampling speed : 100 KHZ

Data transfer : DMA, Interrupt, software

On board memory : 1K

Resolution : 12 bits

Input range : ± 5 v, ± 10 v

Accuracy : ± 12 v

Engine Dynamometer

Make : Siemens Schuchertwerke AG,Germany

Type : Swinging field electric 06992-4, DC Gen/ Mot

Excitation : 220 ... 75 V and 24-48 A and 6500 rpm

Volts Amps Rpm KW

220 125 – 122 4300 - 6000 30

47 – 185 133 600 – 300 44 – 22

185 133 – 136 3000 – 6000 22

105

Sampling Pump

Make : Bosch

Model : ETD 020.00

Suction volume : 0.330 * 10 " m3

Piston travel time approximately : 2 s

Diameter of the sooted surface : 0.30 cm

Permissible range of the pick up : 500 o

C

Smoke Meter

Make : Bosch

Model : ETD 020.50

Supply voltage : 4.5 volts

Measuring Range : 0- 10 BSU

Lamp : 3.8 V 10.07 A

Chemiluminiscent Analyzer for NOx measurement

Make : Rosemount Analytical

Model : 951 A

Ranges : selectable full scale range of

10,25,100,250,2000,2500, and

10000 parts per million

Sensitivity : 0.1 ppm on 10 ppm range

Linearity : ±1% of full scale

Response Time : approximately one second

on all ranges except 10 ppm

Precision : ±5% of full scale

Stability Zero : 1% of full scale in 24 hours

Span : 1% of full scale in 24 hours

Detector operating : ambient

temperature

Recorder output : selectable output of 10 millivolts,

100 millivolts and 1 volt of 5 volt

Ambient temperature : 4.4 C to 37.7 C

Electric power requirements : 107 to 127 VAC, 50/60 Hz,

1000watts

Weight : 10 kg

106

CHAPTER 7

CONCLUSIONS AND SCOPE FOR FUTURE WORK

CONCLUSIONS

The following important conclusions are drawn from the present investigations

concerning fuel additive and in-cylinder turbulence effects on a direct injection diesel

engine.


The use of polymer base additives in different proportions show a significant

influence on diesel engine performance, combustion and emissions characteristics. It is

observed that certain combinations of additives investigated in the present work provides

i. a slightly 'higher ignition delay but a significant reduction in combustion duration;

ii. a smoother combustion, as evident from the lower values of peak pressure and its

derivatives;

iii. a higher rate of heat release in diffusion mixing controlled combustion phase

causing a significant reduction in exhaust smoke (about 37% ) with a marginal

increase in NO (about 4%) concentration.

The additive I at 0.5% addition by volume in diesel fuel yields a maximum

increase in brake specific fuel consumption (BSFC) of 7.6% with a decrease in exhaust

smoke level of 36.8% and an increase of 3.8% in NO concentration.

107

7.2 IN-CYLINDER TURBULANCE INDUCEMENT

In the present investigation, the turbulence inducement inside the engine cylinder

is achieved through use of horizontal bluff bodies and internal jets. The experimental

results obtained using the different arrangements investigated in the present work show

that

i. The engine performance is affected by the presence of bluff bodies in the engine

cylinder through turbulence generation effects and heat loss effects;

ii. The horizontal bluff bodies, in general, result in inferior fuel economy and exhaust

smoke. However, in presence of bluff bodies, there is some improvement in NOx

concentration, primarily due to the consideration of the heat loss;

iii. The orientations of the horizontal rods placed across piston cavity have an

influence on the engine performance and emissions;

iv. At low loads, the horizontal bluff bodies provide a significant decrease in smoke

levels, particularly in the case of rods placed parallel to the piston;

v. The turbulence induced due to internal jets is superior to that induced by bluff

bodies e.g. improvements of about 8 percent in brake specific fuel consumption

(BSFC) and about 20 percent in exhaust smoke over the base values of the engine

at full load are observed. There is however, an increase in NOx (ranging between

10-15%) concentrations over entire load range except at full load condition;

vi. All the configurations of internal jets resulted in higher turbulence and lead to

more prominent recirculating regions near the bowl edge compared to the base

configuration.

vii. The combined effects of best fuel additive case and the best internal jet

configuration provide an increase, under full load conditions, in brake thermal

108

efficiency of about 9.8% in brake specific fuel consumption (BSFC) with a

significant reduction in smoke level of about 38.5%, there is however a marginal

increase of about 4.5% in the exhaust NO as compared to base engine value at full

load condition.

In general, the arrangement of two hole internal jets is found to be superior in

terms of brake thermal efficiency, smoothness of combustion and exhaust smoke

improvements over the other configurations of internal jets. If this system is used

in conjunction with other techniques of in-cylinder turbulence inducement and/or

superior fuel changes, there is a possibility of enhancing the existing performance

and emissions characteristics of a diesel engine as demonstrated through a

representative experimentation carried out using combined changes involving fuel

additive and fuel-air mixing considerations.

7.3 SCOPE FOR FUTURE WORK

i. The effect of the present additive can be tested In automotive engine and in

conjunction with many alternative fuels in use.

ii. The effects of the various shapes of bluff body arid the internal jets can be

investigated.

iii. Several other arrangement of introducing internal jets can be tried in conjunction

with other possible turbulence inducement methods.

iv. A more detailed CFD analysis for reactive conditions needs to be undertaken for

more elaborate understanding and explanation of the results.

109

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LIST OF PAPERS PUBLISHED BASED ON THIS THESIS

I. NATIONAL CONFERENCE

1. R. Venkatesh Babu, S. Sendilvelan.(2008) Effect Of Fuel Additives On The

Formation Of Carbon During Combustion. Proceedings of Recent trends in

Automobile Engineering, Chennai.

2. R. Venkatesh Babu, S. Sendilvelan.(2008) The influence of ethanol blended diesel

fuels on emissions from a diesel engine. Proceedings of Recent trends in

Automobile Engineering, Chennai.

II. INTERNATIONAL CONFERENCE

1. R. Venkatesh Babu, S. Sendilvelan.(2008) Studies On The Effects Of Turbulence

Inducement On Diesel Engine Combustion. Fifth International Conference on

Mechanical Engineering (ICME 2008), Germany. (Communicated)

III. NATIONAL JOURNAL

1. R. Venkatesh Babu, S. Sendilvelan.(2008) Investigations On The Effects Of

Turbulence Inducement On Diesel Engine Combustion. Journal of IIPE, India.

(Communicated)

IV. INTERNATIONAL JOURNAL

1. R. Venkatesh Babu, S. Sendilvelan.(2008) The Effects Of Turbulence Inducement

On Diesel Engine Combustion And Emission Characteristics. Journal of the Brazilian

Society of Mechanical Sciences (Communicated)

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