FACULTY OF ENGINEERING & TECHNOLOGY (AUTOMOBILE...
Transcript of FACULTY OF ENGINEERING & TECHNOLOGY (AUTOMOBILE...
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
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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.
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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
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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.
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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.
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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
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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 …………………………………………………………
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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
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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
6.2 Variation of BSFC with load (additive - 2) ……………………………... 42
6.3 Variation of BSFC with load (additive - 3) ……………………………... 43
6.4 Variation of BSFC with load (additive - 4) ……………………………... 43
6.5 Variation of BSFC with load (additive - 5) ……………………………... 44
6.6 Variation of BSFC with load (additive - 6) ……………………………... 44
6.7 Variation of smoke with load (additive - 1) …………………………….. 45
6.8 Variation of smoke with load (additive - 2) …………………………….. 45
6.9 Variation of smoke with load (additive - 3) …………………………….. 46
6.10 Variation of smoke with load (additive - 4) …………………………….. 46
6.11 Variation of smoke with load (additive - 5) …………………………….. 47
6.12 Variation of smoke with load (additive - 6) …………………………….. 47
6.13 Variation of NO with load (additive - 1) ………………………………... 49
6.14 Variation of NO with load (additive - 2) ………………………………... 49
6.15 Variation of NO with load (additive - 3) ………………………………... 50
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6.16 Variation of NO with load (additive - 4) ……………………………….. 50
6.17 Variation of NO with load (additive - 5) ………………………………... 51
6.18 Variation of NO with load (additive - 6) ………………………………... 51
6.19 Variation of cylinder gas temperature with crank angle (additive - 1) …. 52
6.20 Variation of cylinder gas temperature with crank angle (additive - 2) …. 52
6.21 Variation of cylinder gas temperature with crank angle (additive - 3) …. 53
6.22 Variation of cylinder gas temperature with crank angle (additive - 4) …. 53
6.23 Variation of cylinder gas temperature with crank angle (additive - 5) …. 54
6.24 Variation of cylinder gas temperature with crank angle (additive - 6) …. 54
6.25 Variation of IMEP with load (additive - 1) ……………………………... 55
6.26 Variation of IMEP with load (additive - 2) ……………………………... 55
6.27 Variation of IMEP with load (additive - 3) ……………………………... 56
6.28 Variation of IMEP with load (additive - 4) ……………………………... 56
6.29 Variation of IMEP with load (additive - 5) ……………………………... 57
6.30 Variation of IMEP with load (additive - 6) ……………………………... 57
6.31 Variation of peak cylinder temperature with load (additive-1) …………. 59
6.32 Variation of peak cylinder temperature with load (additive-2) …………. 59
6.33 Variation of peak cylinder temperature with load (additive-3) …………. 60
6.34 Variation of peak cylinder temperature with load (additive-4) …………. 60
6.35 Variation of peak cylinder temperature with load (additive-5) …………. 61
6.36 Variation of peak cylinder temperature with load (additive-6) …………. 61
6.37 Variation of total combustion duration with load (additive-1) …………. 62
6.38 Variation of total combustion duration with load (additive-2) …………. 62
6.39 Variation of total combustion duration with load (additive-3) …………. 63
6.40 Variation of total combustion duration with load (additive-4) …………. 63
6.41 Variation of total combustion duration with load (additive-5) …………. 64
6.42 Variation of total combustion duration with load (additive-6) …………. 64
6.43 Variation of exhaust gas temperature with load (additive-1) …………… 65
6.44 Variation of exhaust gas temperature with load (additive-2) …………… 65
6.45 Variation of exhaust gas temperature with load (additive-3) …………… 66
6.46 Variation of exhaust gas temperature with load (additive-4) …………… 66
6.47 Variation of exhaust gas temperature with load (additive-5) …………… 67
6.48 Variation of exhaust gas temperature with load (additive-6) …………… 67
6.49 Variation of cylinder wall temperature with load (additive-1) ………… 68
6.50 Variation of cylinder wall temperature with load (additive-2) ………… 68
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6.51 Variation of cylinder wall temperature with load (additive-3) ………… 69
6.52 Variation of cylinder wall temperature with load (additive-4) ………… 69
6.53 Variation of cylinder wall temperature with load (additive-5) ………… 70
6.54 Variation of cylinder wall temperature with load (additive-6) ………… 70
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
6.58 Variation of BSFC with percentage additive (additive-4) ……………... 72
6.59 Variation of BSFC with percentage additive (additive-5) ……………... 73
6.60 Variation of BSFC with percentage additive (additive-6) ……………... 73
6.61 Variation of smoke with percentage additive (additive-1) ……………... 74
6.62 Variation of smoke with percentage additive (additive-2) ……………... 74
6.63 Variation of smoke with percentage additive (additive-3) ……………... 75
6.64 Variation of smoke with percentage additive (additive-4) ……………... 75
6.65 Variation of smoke with percentage additive (additive-5) ……………... 76
6.66 Variation of smoke with percentage additive (additive-6) ……………... 76
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
6.77 Variation of peak pressure with load ……………………………………. 91
6.78 Variation of maximum rate of pressure rise with load ………………….. 92
6.79 Variation of ignition delay with load …………………………………… 92
6.80 Variation of combustion duration with load ……………………………. 93
6.81 Variation of smoke number with load …………………………………... 93
6.82 Variation of nitric oxide with load ……………………………………… 94
6.83 Variation of brake thermal efficiency with load ………………………... 94
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6.84 Variation of peak pressure with load ……………………………………. 95
6.85 Variation of maximum rate of pressure rise with load ………………….. 95
6.86 Variation of ignition delay with load …………………………………… 96
6.87 Variation of combustion duration with load ……………………………. 96
6.88 Variation of smoke number with load …………………………………... 97
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
the combined case with the base engine ………………………………… 100
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
the combined case with the base engine ………………………………… 102
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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
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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.
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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
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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
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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-
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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.
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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
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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.
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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.
6.1 ADDITIVE WITH FUEL
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
Fig.6.2 Variation of BSFC with load (additive - 2)
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
Fig.6.3 Variation of BSFC with load (additive - 3)
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
Fig.6.4 Variation of BSFC with load (additive - 4)
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
Fig.6.5 Variation of BSFC with load (additive - 5)
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.6.6 Variation of BSFC with load (additive - 6)
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
Fig.6.8 Variation of smoke with load (additive - 2)
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
Fig.6.9 Variation of smoke with load (additive - 3)
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
Fig.6.10 Variation of smoke with load (additive - 4)
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
Fig.6.11 Variation of smoke with load (additive - 5)
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
Fig.6.12 Variation of smoke with load (additive - 6)
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
Fig.6.14 Variation of NO with load (additive - 2)
50
100
200
300
400
500
600
700
800
0 20 40 60 80 100LOAD(%)
NO
-pp
m
BASE a105
a110 a115
a120
Fig.6.15 Variation of NO with load (additive - 3)
100
200
300
400
500
600
700
800
0 20 40 60 80 100LOAD(%)
NO
-ppm
BASE a105
a110 a115
a120
Fig.6.16 Variation of NO with load (additive - 4)
51
100
200
300
400
500
600
700
800
0 20 40 60 80 100
LOAD(%)
NO
-pp
m
BASE a105
a110 a115
a120
Fig.6.17 Variation of NO with load (additive - 5)
100
200
300
400
500
600
700
800
0 20 40 60 80 100
LOAD(%)
NO
-pp
m
BASE a105
a110 a115
a120
Fig.6.18 Variation of NO with load (additive - 6)
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
Fig.6.20 Variation of cylinder gas temperature with crank angle (additive - 2)
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
Fig.6.21 Variation of cylinder gas temperature with crank angle (additive - 3)
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
Fig.6.22 Variation of cylinder gas temperature with crank angle (additive - 4)
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
Fig.6.23 Variation of cylinder gas temperature with crank angle (additive - 5)
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
Fig.6.24 Variation of cylinder gas temperature with crank angle (additive - 6)
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
Fig.6.26 Variation of IMEP with load (additive - 2)
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
Fig.6.27 Variation of IMEP with load (additive - 3)
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
Fig.6.28 Variation of IMEP with load (additive - 4)
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
Fig.6.29 Variation of IMEP with load (additive - 5)
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
Fig.6.30 Variation of IMEP with load (additive - 6)
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
Fig.6.32 Variation of peak cylinder temperature with load (additive-2)
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
Fig.6.33 Variation of peak cylinder temperature with load (additive-3)
500
1000
1500
2000
2500
0 20 40 60 80 100
LOAD(%)
Tm
ax-K
0.005 0.01
0.015 0.02
BASE
Fig.6.34 Variation of peak cylinder temperature with load (additive-4)
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
Fig.6.35 Variation of peak cylinder temperature with load (additive-5)
500
1000
1500
2000
2500
0 20 40 60 80 100
LOAD(%)
Tm
ax-K
0.005 0.01
0.015 0.02
BASE
Fig.6.36 Variation of peak cylinder temperature with load (additive-6)
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
Fig.6.38 Variation of total combustion duration with load (additive-2)
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
Fig.6.39 Variation of total combustion duration with load (additive-3)
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
Fig.6.40 Variation of total combustion duration with load (additive-4)
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
Fig.6.41 Variation of total combustion duration with load (additive-5)
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
Fig.6.42 Variation of total combustion duration with load (additive-6)
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
Fig.6.44 Variation of exhaust gas temperature with load (additive-2)
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
Fig.6.45 Variation of exhaust gas temperature with load (additive-3)
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.6.46 Variation of exhaust gas temperature with load (additive-4)
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
Fig.6.47 Variation of exhaust gas temperature with load (additive-5)
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
Fig.6.48 Variation of exhaust gas temperature with load (additive-6)
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
Fig.6.50 Variation of cylinder wall temperature with load (additive-2)
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
Fig.6.51 Variation of cylinder wall temperature with load (additive-3)
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.6.52 Variation of cylinder wall temperature with load (additive-4)
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
Fig.6.53 Variation of cylinder wall temperature with load (additive-5)
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.6.54 Variation of cylinder wall temperature with load (additive-6)
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
Fig.6.56 Variation of BSFC with percentage additive (additive-2)
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
Fig.6.57 Variation of BSFC with percentage additive (additive-3)
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
Fig.6.58 Variation of BSFC with percentage additive (additive-4)
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
Fig.6.59 Variation of BSFC with percentage additive (additive-5)
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%
Fig.6.60 Variation of BSFC with percentage additive (additive-6)
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
Fig.6.62 Variation of smoke with percentage additive (additive-2)
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
Fig.6.63 Variation of smoke with percentage additive (additive-3)
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
Fig.6.64 Variation of smoke with percentage additive (additive-4)
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
Fig.6.65 Variation of smoke with percentage additive (additive-5)
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
Fig.6.66 Variation of smoke with percentage additive (additive-6)
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
combined case with the base engine
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
combined case with the base engine
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
combined case with the base engine
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.
7.1 ADDITIVE WITH FUEL
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)