Accepted Manuscript Effects of Combustion Chamber Geometry on Combustion Characteristics of a DI Diesel Engine Fueled with Calophyllum Inophyllum Methyl Ester Dr. B.R. Ramesh Bapu, Dean R&D, L. Saravanakumar, Research Scholar, Dr. B. Durga Prasad, Professor PII:
S1743-9671(15)20599-5
DOI:
10.1016/j.joei.2015.10.004
Reference:
JOEI 184
To appear in:
Journal of the Energy Institute
Received Date: 12 March 2015 Revised Date:
7 October 2015
Accepted Date: 14 October 2015
Please cite this article as: B.R. Ramesh Bapu, L. Saravanakumar, B. Durga Prasad, Effects of Combustion Chamber Geometry on Combustion Characteristics of a DI Diesel Engine Fueled with Calophyllum Inophyllum Methyl Ester, Journal of the Energy Institute (2015), doi: 10.1016/ j.joei.2015.10.004. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
ACCEPTED MANUSCRIPT Effects of Combustion Chamber Geometry on Combustion Characteristics of a DI Diesel Engine Fueled with Calophyllum Inophyllum Methyl Ester Dr.B.R.Ramesh Bapua, L.Saravanakumar*,b, Dr.B.Durga Prasadc a
Dean R&D, Chennai Institute of Technology, Chennai, India.
*,bResearch Scholar, Dept. of Mechanical Engg., JNTU, Anantapuramu, India. Professor, Dept. of Mechanical Engg., JNTU college of Engg., JNTU, Anantapuramu, India. a
RI PT
c
[email protected], * b
[email protected],
[email protected]
Abstract
This paper describes the results of an experimental investigation carried out in a single
SC
cylinder, variable compression ratio, CI engine fuelled with Calophyllum Inophyllum Methyl Ester (CIME) blended with diesel. An earlier investigation made by the investigators using
M AN U
CIME blends in conventional Hemispherical Combustion Chamber (HCC), showed that the blend B20 provides the optimum results and hence B20 blend was used as a test fuel for further investigations. Further, the emissions like unburned hydrocarbon (UBHC), carbon monoxide (CO) and smoke have been noticed higher with diesel. In this investigation, attempts have been made to reduce the emissions and improve the combustion characteristics by enhancing the fuel-air mixture preparation and its turbulence by changing the design of
TE D
piston bowl geometry. For this, a Modified Hemispherical Combustion Chamber (MHCC) has employed and the results were compared with conventional Hemispherical Combustion Chamber (HCC). The fuel-air mixture formation in the cylinder was simulated at different positions of the piston (at TDC, mid of stroke and at BDC) using Ansys Fluent software.
EP
Conclusively, from the investigations, MHCC was recognized as an ideal choice of combustion chamber design for the entire range of operations of the engine using the blend
AC C
(B20) than that of HCC.
Key words: Biodiesel; Calophyllum Inophyllum Methyl Ester; Combustion; Compression ratio; combustion chamber geometry; emissions. 1. Introduction
Most of the investigators believe that green fuels, widely known as biofuels are more environment friendly than the prevalent fossil fuels. The emissions from the diesel engines pollute the atmosphere and cause various ill-effects. Besides, green fuels are a strong contender in the race for the fuels to replace fossil fuels. Nowadays, researchers are focusing especially on emissions from diesel engines, and are trying to implement key technologies to
ACCEPTED MANUSCRIPT meet the future emission norms. The automotive sectors have been identified as one of the major contributors to atmospheric pollution. The emission issues are putting great pressure on automotive sectors to develop new technologies, which can reduce the emissions and in turn improves the efficiency in a better manner. Further, it progresses steadily towards the reduction of harmful emissions by looking for alternate fuels with minor modifications in the
RI PT
engine configurations. However, there is still room to achieve higher thermal efficiencies and lower emissions.
One of the methods to accomplish better fuel-air mixture formation within the combustion chamber by redesigning the combustion chamber to improve the thermal efficiency and
SC
reduce the emission levels. In this investigation, change in combustion chamber design is accomplished by modifying the piston bowl geometry and the design parameters are
M AN U
optimized by Ansys fluent software. Experimental investigations carried out by several researchers on the effect of piston bowl geometries on diesel engines fuelled with biodiesel are discussed herein, hereafter.
Yang et al.[1] studied the effects of piston bowl geometry on combustion and emission characteristics of a diesel engine using biodiesel. Three combustion geometries, namely HCC
TE D
(hemispherical combustion chamber), SCC (shallow depth combustion chamber) and OCC (omega combustion chamber) were taken for the experiment. They reported that, narrow entrance of combustion chamber results in better squish and enhances the air-fuel mixture at higher engine speeds than SCC. However, OCC geometry dominates in combustion process
EP
by generating strong squish, better mixture formation in short time which shows better performance from medium to high engine speed, and SCC generates higher NO than other two combustion chambers at low engine speed. In an attempt to overcome some of the
AC C
drawbacks encountered in the standard hemispherical open combustion chamber, Jaichandar et al.[2] have successfully studied the influence of reentrant combustion chamber geometry on a diesel engine fueled with Pongamia Oil Methyl Ester (POME).
Two modified
combustion chamber such as Toroidal Reentrant combustion chamber (TRCC) and Shallow depth reentrant combustion chamber (SCC) fueled with B20 and diesel were employed and the results were compared with petroleum based diesel fuels (PBDF). It is reported that higher brake thermal efficiency, high peak pressure and lower specific fuel consumption, ignition delay were occurred while using 20% blend in TRCC combustion geometry. It is also revealed that sharp reduction of particulate matter (PM), carbon monoxides (CO) and
ACCEPTED MANUSCRIPT unburned hydrocarbons (UBHC) were noticed with TRCC.
However, higher oxides of
nitrogen emissions (NOx) were observed for TRCC. Ravikrishna et.al [3] investigated the effect of induced swirl in piston bowl used in diesel engines. They found that the reduction in emission has occurred by the introduction of high swirl reentrant bowl geometry with sac-less injector. Further, they used CFD
RI PT
simulation analysis to investigate the combustion and emission parameters, thereby modified geometry with low central projection on bowl with low sac-volume injector yields better combustion and lower emissions. An injection timing of 8.6° CA bTDC was found to be optimum since it led to a 27% reduction in NOX emissions and 85% reduction in
SC
soot levels as compared to the baseline configuration. In an attempt to optimize the combustion chamber geometry, Jaichander et al. [4] investigated the effect of varying the
M AN U
combustion chamber geometry on the performance of a CI engine using Pongamia biodiesel. The experiments were carried out using a blend of 20% pongamia oil biodiesel with diesel on three different types of combustion chamber geometries such as Hemispherical(HCC), shallow depth (SCC) and Toroidal combustion chamber(TCC) at a given compression ratio. The results show that, for B20 blend with toroidal chamber shown increased brake thermal efficiency significantly and lower the specific fuel consumption compared to SCC & HCC
TE D
due to better air motion. Taghavifar et al.[5] conducted a test on CI engine to investigate role of piston bowl geometries on flow behavior, combustion and performance characteristics. It is reported that homogeneity factor and equivalence ratio of air-fuel mixing process, combustion initiation during heat release rate and pressure curves are the two major
EP
parameters, which plays a significant role on combustion. However, smaller bowl size induces better squish and vortex formation with lesser spray penetration and reduced ignition
AC C
delay.
Shengli Wei et al.[6] carried out numerical analysis on the effect of swirl ratios on swirl chamber combustion system of DI diesel engines. In order to improve the spatial spray distribution and enhanced airflow movement in a combustion chamber, a new swirl chamber has been proposed. They reported that after spray, the temperature distribution and kinetic energy of the in-cylinder mixture is influenced by the combustion swirl and squish. The swirl ratios of 0.8 and at 2.7, the NO mass fraction fluctuates between lowest and highest respectively, and with swirl ratios of 0.2 and 3.2, the Soot mass fraction lies between lower and higher percentages respectively. However, the swirl ratio of 0.2 is better in view of fuelair equivalence ratio and swirl ratio of 0.8 is concluded as the suitable combustion chamber to
ACCEPTED MANUSCRIPT provide better performance and combustion. Ramesh Mamilla et al.[7] made an investigation on a single cylinder, air-cooled, four stroke diesel engine with different Jatropha methyl ester-diesel blends, by varying the piston bowl geometry.
Three different piston bowl
geometries were employed such as Spherical, Toroidal and Re-entrant and the effect of these geometries on performance, combustion and emission characteristics were studied. It was
RI PT
found that the 20% blend provided better performance and lower emissions like HC, CO and smoke with toroidal chamber. However, nitrogen oxides were significantly lower than other two geometries but higher when compared to conventional diesel in spherical chamber.
Venkateswaran et al.[8] investigated the effect of re-entrant bowl geometry on performance
SC
and combustion in a direct injection turbocharged diesel engine. They reported that, piston bowl geometry influences much on the performance and combustion of a diesel engine. The
M AN U
bowl geometry and dimensions such as bowl lip area, pip region and toroidal radius are all enhances the diffusion combustion at the later combustion stage.
The parameters like
compression ratio squish clearance, bowl diameter, injection rate were kept constant, and three different bowl geometries were considered for the investigation. The bowl geometry designs were simulated with CFD STAR-CD software and modeled the in cylinder flow process and it has been compared with the experimental results. Results showed that bowl 3
TE D
enhances turbulence and better air-fuel mixing, which results in reduced indicated specific fuel consumption and soot emission at the cost of rise in NOx due to better mixing and fast combustion. The investigation carried out by Bora et al. [9] using biodiesel obtained from the mixture of polanga oil, karanja oil and jatropha oil in equal proportion in CI
EP
engines and compares the performance and emission analysis with diesel fuel. The BTE has improved significantly up to 40% of the mixture blend with diesel and beyond
AC C
which thermal efficiency decreases due to owing its high viscosity and poor volatility inturn reduces the combustion efficiency.
A.S.Ramadhas et al.[10] developed a
thermodynamic model to analyze the performance characteristics of CI engine, such a way that it can be used for characterizing any hydrocarbon fueled engine (diesel, biodiesel and their blends). It was reported that for the rise in compression ratio, peak pressure and peak temperature, Brake Thermal Efficiency was increased and the above-mentioned factors are decreased with increase in relative air fuel ratio. Venkanna [11] conducted a test in a DI diesel engine using Calophyllum Inophyllum linn oil (hone oil) methyl esters (100% biodiesel) to investigate the performance, combustion and emission characteristics. He
ACCEPTED MANUSCRIPT found that at high injector opening pressure exhibits increased brake thermal efficiency and reduced CO, HC and smoke emissions. This investigation aims to achieve better combustion with modified piston bowl design using biodiesel (20%)-diesel blend. The effect of engine combustion parameters such as in cylinder pressure variations with crank angle, rate of peak pressure rise, net heat release rate, ignition
RI PT
delay and emission characteristics were discussed comprehensively by varying the compression ratio at different piston bowl geometries. The test results are compared with diesel in conventional piston (HCC) bowl.
2.1 Materials – Biodiesel production – Test fuels
SC
2. Materials and Methods
M AN U
The apparatus used for transesterification consists of a 5 litre four necked round bottom flask. One of the four equipped with thermo well for inserting a platinum RTD temperature sensor with an accuracy of ±1°C interfaced to a digital indicator, one for fixing condenser setup, third one for pouring the raw oil and the center has been equipped with the stirrer with motor arrangement, and the regulator controlled the speed of the motor. The CI oil is heated to about 120°C for 30 min to remove moisture and is allowed to cool. One litre of moisture free
TE D
CI oil is used for esterification process. Since raw CI oil having high amount of free fatty acids and has the acid value of 39.6 mg KOH/g of oil, makes difficulty in the process of alkaline transesterification because of soap formation. Hence, the crude CI oil requires two
EP
stage esterification process. First stage, namely acid esterification, requires methanol and concentrated sulphuric acid as a catalyst to carry out the process. The mixture of CI oil, methanol and catalyst is heated at the constant temperature and constantly stirred during the
AC C
process. This process converts the free fatty acids into esters by reducing the free fatty acid content to less than 3% to facilitate alkaline process. In second stage, transesterification process, the bottom product from the first stage is mixed with methanol and catalyst KOH in right proportion along with constant heating and stirring to obtain the biodiesel (CIME). A separating funnel is used for collection of the final product. Calophyllum Inophyllum Methyl Ester (20%) blended with diesel (80%) called as B20 blend and diesel is used in this test. The fatty acid profile of raw Calophyllum Inophyllum oil and physio-chemical properties of biodiesel blend and diesel are presented in Table1 and Table2.
ACCEPTED MANUSCRIPT
Table 1. Fatty acids profile* Fatty Acids
Values
Palmitic
14.8-18.5
Stearic
6.1-19.2
Oleic
36.2-53.1
Linolenic
---
Linoleic
15.8-28.5
Arachidic
--Erucic 3.3
SC
Any special
RI PT
Composition
fatty acid
Table 2. Physio-chemical properties Parameters
Protocol
ASTM D4809
Flash point (°C)
ASTM D93
Kinematic Viscosity (CSt)
ASTM D445
---
130 min
Values
B20
B100
42380
38900 43600
68
1.9 to 6.0
EP
Density(kg/m3)
Biodiesel ASTM limits
TE D
Heating value (kJ/kg)
M AN U
* Source from [12]
3.5
Diesel
148
57
4.3
2.98
841
861
836
Acid value (mg KOH/g)
---
---
1.63
---
Cloud point (°C)
ASTM D2500
---
7.9
13.2
6.5
Pour point (°C)
ASTM D97
-15 to 10
2.9
4.3
3.1
Specific gravity@15°C
ASTM D4052
0.87 to 0.89
0.85
0.86
0.85
Cetane Number
ASTM D613
48 to 70
52
57
50
AC C
---
ACCEPTED MANUSCRIPT 2.2 Engine Setup A direct injection, naturally aspirated, water-cooled, kirloskar (model TV-1), single cylinder, 4-stroke variable compression ratio CI engine, is used in this investigation. The test engine has a displacement volume of 661 cc and its compression ratio can be varied between 12 to 18:1. The engine is directly coupled to an eddy current dynamometer for loading. The
RI PT
load on the engine is measured with a load cell (S beam, universal, strain gauge type) which is mounted on the dynamometer cradle arm. It is capable of developing 3.5 kW at the rated speed 1500 rpm. The engine is equipped with Bosch injection pump and multi hole injector (3 hole x 0.24 mm dia.). The standard piston has a hemispherical combustion chamber with a
SC
volume of 32cc. The start of injection and injector opening pressure are 23° bTDC (before
2.3 Instrumentation system
M AN U
top dead centre) and 220 bar as per the manufacturer setting.
The intake air flow rate is measured with the help of an orifice meter, fuel flow rate with flow transmitter, cooling water and calorimeter water flow with rotameter.
An inductive pickup
sensor with a resolution of 1 degree was used to measure the engine speed. The measurement of cylinder pressure was made using a Quartz (piezo-electric) dynamic pressure transducer
TE D
with built in amplifier (model M111A22, PCB Piezotronics, INC.) having a maximum pressure 1034 bar (5000 psi) with a resolution of 0.00068 bar. The transducer, mounted on a suitable adaptor, was made to communicate with the combustion chamber through a communicating passage, drilled into the cylinder head. The charge output of the transducer,
EP
which is proportional to the rate of pressure difference with time was fed to a charge amplifier to provide the voltage signal proportional to the cylinder pressure. The five gas
AC C
analyzer (model NPM-MGA-1, NETEL INDIA LTD.) has been involved to determine the concentration of gaseous pollutant species from the engine exhaust tail pipe. A probe from the analyzer unit is inserted into the exhaust tail pipe of the engine to measure the emission concentration while the engine running. The smoke meter (model AVL 437 C, serial number 02766) is employed for the measurement of smoke. The schematic diagram of the experimental setup is given in Figure1. The technical specification of the VCR engine used in this study has shown in Table 3.
ACCEPTED MANUSCRIPT
SC
RI PT
Table 3. Engine Specifications Particulars Specifications Engine make Kirloskar, Model TV1 No. of cylinder Single No. of stroke Four stroke Bore and Stroke 87.5 mm and 110 mm Power & speed 3.5 kW & 1500 Piezo sensor Range 5000 Psi 23°BTDC Fuel injection timing Valve timing Inlet valve Opening 4.5° BTDC Inlet valve closing 35.5° ATDC 35.5° BTDC Exhaust valve opening 4.5° ATDC Exhaust valve closing Injector opening pressure 220 bar
DIESEL
BIODIESEL TANK
TANK
FUEL FLOW TRANSMITTER
M AN U
MONINTOR
DATA AQUISITION SYSTEM
MULTIGAS ANALYZER
EP
SMOKE METER
TE D
CHARGE AMPLIFIER
AIR BOX CRANK ANGLE SENSOR EDDY CURRENT DYNAMOMETER
AC C
ENGINE
AIR FLOW TRANSMITTER
Fig. 1 Schematic diagram of experimental setup
The engine runs at constant speed of 1500 rpm and the observations have recorded for the variations of load from 0% to 100% at an increment of 25%. The performance characteristics and pressure-crank angle data for every degree crank angle from 0° to 360° for each cycle
ACCEPTED MANUSCRIPT was recorded. This experiment was repeated using standard piston and modified piston shown in Figure 2 and 3 and the results were compared with each other and presented. 2.4 Piston geometry description In this study, two different piston bowl geometries were considered which are shown in Figure 2 and 3. These geometries usually employed to optimize the combustion process in
RI PT
the diesel engines. Figure 3 is standard hemi-spherical combustion chamber (HCC) with standard aspect ratio. (i.e., diameter to depth ratio) and Figure 2 is modified hemi-spherical combustion chamber (MHCC) with lower diameter/depth ratio. Figure 4 and 5 shows the top
TE D
M AN U
SC
view of MHCC and HCC geometry.
Fig. 3 Standard piston (HCC)
AC C
EP
Fig. 2 Modified piston (MHCC)
Fig. 4 Modified piston (wireframe top view)
Fig. 5 Standard piston (wireframe top view)
ACCEPTED MANUSCRIPT 3. Results and discussion 3.1 Theoretical considerations The comparative study of flow simulation and experimental results of the standard piston with modified piston were performed. The simulation was done using ANSYS fluent
RI PT
software, the experiments were performed on a VCR engine, and the results are compared and presented. The velocity profile of fluid at the interior of the model was generated for various positions of the piston such as (TDC) top dead centre, mid of stroke and (BDC) bottom dead centre shown in Fig. 8 to Fig. 13 for both the standard and the modified pistons.
Fig. 7 meshed view of modified piston
AC C
EP
TE D
Fig. 6 meshed view of standard piston
M AN U
SC
In addition, the meshed views are shown in Figure 6 and 7.
Fig. 8 Standard piston at BDC (Front)
Fig. 10 Standard piston during travel (Front)
Fig. 9 Modified piston at BDC (Front)
Fig. 11 Modified piston during travel (Front)
RI PT
ACCEPTED MANUSCRIPT
Fig. 12 Standard piston at TDC (Front)
Fig. 13 Modified piston at TDC (Front)
Figures 8 to 13 show the variations of fluid velocities at different θ values for both standard and
SC
modified piston. On comparing the inlet velocity for both standard and modified pistons, significant increase in turbulence occurs in the modified piston and it is continued till the half
M AN U
the way of its stroke. There is noticeable increase in velocity occurs at TDC position in the modified piston. It may be due to the squish motion of the gas mixture inside the cylinder. Squish [13] is the gas motion that travels radially inwards or transverse and axially mixed near the TDC, results better in charge preparation, burning rate and heat release rate may increase. The theoretical squish velocity can be calculated from the relation given by Heywood [13] in
and heat transfer.
=
− 1
TE D
equation (1) by neglecting the effects of gas dynamics, friction, leakage past the piston rings,
(1)
EP
where VB = volume of the piston bowl, m3; Ac=cross sectional area of the cylinder, m2; Sp=instantaneous piston speed, rpm; z= distance between the piston crown and the cylinder
AC C
head, m;(z = c + Z) Z = l + a – S; S = stroke length, m; l = connecting rod length, m; a = crank radius, m;
It is observed that the maximum squish velocity occurs at 12° bTDC and the increase in squish velocity with modified piston is around 35% compared with standard piston. The CFD analysis reveals that, the increase in squish velocity increases the charge motion characteristics in a modified piston engine.
ACCEPTED MANUSCRIPT 3.2 Combustion Analysis The combustion and emission characteristics of a Variable Compression Ratio CI engine were analyzed using the CIME blend B20 to investigate the effect of piston bowl geometry and the results are compared with diesel. The combustion characteristics such as pressure-
discussed. 3.2.1 Rate of Pressure Rise and Peak Pressure
RI PT
crank angle, heat release rate, ignition delay and rate of pressure rise were investigated and
Figure 14 (a), (b), (c) illustrates the trend in the variation of the rate of maximum pressure
SC
rise for the blend CIME-diesel (B20) and diesel for the different compression ratios and different combustion chamber geometries. It could be seen that, for a given CR, as the load
M AN U
increases the fluctuation in peak pressure rise, increases up to 65% of load. After 65% load, the fluctuation started to decrease as the load increases until the full load. This is because as the load increases, the quantity of fuel injected is more and more fuel is evaporated; therefore, the fluctuation is more due to more irregular and random evaporation. Besides, it can be observed that rate of pressure rise is a function of the intensity of the initial premixed burning of fuels. It is governed by the delay period, spray envelops of the injected fuel and
TE D
the charge temperature, which are appreciably influenced by the blends and more significant for modified combustion chamber geometry.
Compared with diesel operation for the various configurations tried, increasing the CR results
EP
in substantial improvement in the values of peak pressure for the blend B20. However, there is a marginal increase in peak pressure occurs for B20 than diesel using modified geometry. This could be due to decrease in delay period with an increase in CR or by intensive mixing
AC C
of fuel-air mixture. In a conventional bowl (HCC) with diesel, the rate of pressure rise recorded as 4.13 occurring at 8° aTDC and using B20 it was noticed as 3.72 at CR16, 4.08 at CR17, and 4.02 at CR18 occurring at 3°aTDC, 1°aTDC, 7°aTDC, respectively. Likewise, fueling B20 with MHCC, 4.14 for CR16, 4.02 for CR17 and 3.84 for CR18 occurring at 1°aTDC, 9°aTDC, and 5°aTDC respectively. 3.2.2 Pressure-Crank Angle (P-θ) The variations of pressure-crank angle at different bowl geometries and are at different compression ratios under full load conditions are presented in Figure 15 (a), (b), (c). It is noticed that, for the given CR, the peak pressure rise with B20, on modified combustion
ACCEPTED MANUSCRIPT geometry shows significant variations than conventional combustion geometry. It was noticed that peak pressure increases with increase in CR irrespective of the geometry. At higher CR18, the peak pressure values are 54.06 and 57.24, bar occurring at 6°CA aTDC and 5°CA aTDC for HCC and MHCC respectively, compared with diesel in HCC of 54.62 bar occurring at 6°CA aTDC. It is found that, MHCC provides a 4.7% increase in peak pressure
RI PT
than diesel in HCC. This may be due to the large amount of fuel consumed in the uncontrolled phase of combustion, and that is governed by the ignition delay period. For diesel fuel operation, it is noticed that for HCC geometry, an ignition delay of 16°CA, peak pressure of 54.62 bar occurred at 8°aTDC and a peak rate of pressure rise of 4.13bar/°CA. It
SC
can be seen that for geometry MHCC, rise in CR from 16 to 18, a decrease of 23.5% in the delay period was noticed. Further, the peak pressure increased by 4.7% and peak rate of pressure rise decreased by 6.34% compared with diesel in HCC. Besides, it is found that the
3
HCC MHCC HCC-DL
2 1
0 -30 -20 -10 0 -1
10
20
-2
CR18
60
30
40
50
Pressure (bar)
4
70
CR 18
60
TE D
rate of pressure rise, dp/dθ (bar/°CA)
5
M AN U
occurrence of the peak pressure has been advanced by 3°CA than diesel.
-3 Crank angle (degree)
HCC
40 30
MHCC
20
HCC-DL
10 0 -30 -20 -10
0
10
20
AC C
60
70
CR17
4
50 HCC
3
MHCC
2
HCC-DL
1
20
30
40
50
-2 Crank angle (degree)
Fig. 14 (b) RPR Vs crank angle at CR 17
60
Pressure (bar)
rate of pressure rise, dp/dθ (bar/°CA)
50
60
CR17
10
40
Fig. 15 (a) pressure Vs crank angle at CR18
5
0 -30 -20 -10 0 -1
30
Crank angle (degree)
EP
Fig. 14 (a) RPR Vs crank angle at CR 18
50
HCC 40
MHCC
30
HCC-DL
20 10 0
-30 -20 -10
0
10
20
30
40
50
60
Crank angle (degree)
Fig. 15 (b) pressure Vs crank angle at CR17
70
ACCEPTED MANUSCRIPT 60 CR 16
CR16
4
50
3
HCC
2
MHCC
1
HCC-DL
HCC
40
MHCC
30
HCC-DL
20 10
10
20
30
40
50
60 0 -30 -20 -10
-2 Crank angle (degree)
RI PT
0 -30 -20 -10 0 -1
Pressure (bar)
rate of pressure rise, dp/dθ (bar/° CA)
5
0
10
20
30
40
50
60
70
Crank angle (degree)
Fig. 14 (c) RPR Vs crank angle at CR 16
Fig. 15 (c) pressure Vs crank angle at CR16
SC
At full load for various compression ratios (16, 17, and 18) and different geometries, using B20 and diesel, the combustion parameters recorded and shown in Table 4.
HCC
CR
Peak pressure (bar)
16
44.04
17
49.89
18
SOC (° bTDC)
ID (°CA)
44.75
3.72
6
13
41.88
4.08
7
11
35.04
4.02
8
10
46.61
46.5
4.14
6
12
50.31
39.9
4.02
7
10
18
57.24
30.78
3.84
10
8
17.5
54.62
37.12
4.13
7
12
EP
17
AC C
HCC-DL
RPR (bar/°CA)
54.06
16 MHCC
NHR (J/°CA)
TE D
Geometry
M AN U
Table 4. Combustion parameters for various geometries
3.2.3. Heat release rate analysis The net heat release rates for the blend B20 in HCC and MHCC are compared with conventional diesel fuel in HCC at different compression ratio and the variations are illustrated in the Figure16 (a), (b), (c). It is seen that the maximum heat release rate characterized by the peak rate of fuel burning of about 50% occurs at 11° aTDC for B20 than 9° aTDC for diesel. It is also predicted that almost 60 % of total heat supplied have been released at the termination of premixed and diffusion zone and the rest of the heat will be released in the uncontrolled phase of combustion. It is shown in the figure that the shapes of heat release curves at a given CR are nearly identical though differing in magnitude. While
ACCEPTED MANUSCRIPT increasing the compression ratio, the NHR values decreased irrespective of the geometry. It can be seen that at higher CR18, the peak net heat release rate recorded as 35.04 J/°CA for HCC and 30.78 J/°CA for MHCC while comparing diesel at conventional HCC provides 37.12 J/°CA. In comparison with diesel in HCC, the reduction in NHR values is found to be 5.6% for HCC and 17% for MHCC. Moreover, the reduction in combustion duration is high
TE D
EP
Net Heat Release rate (J/°CA)
AC C
Fig. 16 (b) NHR Vs crank angle at CR17 40 35 30 25 20 15 10 5 0 -60 -50 -40 -30 -20 -10 -5 0 10 20 30 -10 Crank angle (degree)
HCC MHCC HCC-DL
40 50 60
Fig. 16 (c) NHR Vs crank angle at CR18
HCC-DL
15 13
CR16
11 9
40 50 60
CR17
MHCC
17
M AN U
HCC-DL
45 HCC 40 35 MHCC 30 HCC-DL 25 20 15 10 5 0 -5 0 10 20 30 40 50 60 -60 -50 -40 -30 -20 -10 -10 Crank Angle (degree)
Neet Het Release rate (J/°CA)
HCC
SC
Ignition delay (°CA)
MHCC
Fig. 16 (a) NHR vs crank angle at CR16
CR18
RI PT
19 HCC
0
0.5
1
1.5
2
2.5
3
3.5
BP (kW)
Fig. 17 (a) Ignition delay vs Brake Power at CR16
Ignition delay (°CA)
50 45 40 35 30 25 20 15 10 5 0 -60 -50 -40 -30 -20 -10 -5 0 10 20 30 -10 Crank Angle (degree) CR16
HCC MHCC HCC-DL
19 17 15 13
CR17
11 9 0
0.5
1
1.5
2
2.5
3
3.5
BP (kW)
Fig. 17 (b) Ignition delay Vs Brake Power at CR17
Ignition delay (°CA)
Net Heat Release rate (J/°CA)
for MHCC followed by HCC.
HCC MHCC
19 17
HCC-DL
15 13 11
CR18
9 7 0
0.5
1
1.5
2
2.5
3
3.5
BP (kW)
Fig. 17 (c) Ignition delay Vs Brake Power at CR18
ACCEPTED MANUSCRIPT For calculation of heat release rate, the single zone model approach has been followed [14]. From the first law of thermodynamics, the energy equation can be written as dQ + ∑m h! = dW + dU
(1)
Neglecting the crevice flow % &' and modify the equation (1) with respect to crank angle θ, (*
−
()+ (*
(0
(1
+ m!,- h. = P (* + (*
RI PT
()
(2)
where U is sensible heat energy of cylinder charge, hf – sensible enthalpy of injected fuel, Q – heat released during combustion, Qw – heat transferred from the cylinder. Since hf is
(*
(0
(0
(6
= P (* + C3 4P (* + V (* 7 +
Using 89 − 8 = :, and
& &;
()+ (*
= < (for diesel k ranges between 1.3 to 1.35) equation (3) can
be modified as (*
=
(0
?
(6
= =>? P (* + =>? V (* +
where ()+ (*
() (*
()+ (*
(4)
= heat release rate, P=Instantaneous pressure, Pa, V=cylinder volume, m3,
TE D
()
(3)
M AN U
()
SC
negligible compared to heat released on combustion and obtain the heat release equation as
= heat transfer, J/deg. With equation (4), the heat transfer rate during the combustion
EP
can be
calculated precisely. The compression and expansion processes of unburned and burned @A
AC C
gases are close to adiabatic isentropic process [13] (pVk = constant, where @ = C). The B
instantaneous pressure can be calculated from the relation pVk = constant
(5)
and cylinder volume V at any crank angle position θ is obtained from [13] V = Vc +
DEF G
(I + J − K)
(6)
where l = connecting rod length, a = crank radius, s = distance between the crank axis and the piston pin axis. The heat transfer across the wall for a diesel engine can be estimated from the Hohenberg’s correlation [10] given as
ACCEPTED MANUSCRIPT ℎ =
?NOPQ.S TUV ?. WQ.S
(7)
Q.QX Y Q.Z
where h= heat transfer coefficient, W/m2K; P= cylinder pressure, bar; Cm=mean piston speed, m/s;
Vc = cylinder volume, m3.
3.2.4. Ignition Delay period
RI PT
Ignition delay [13] is occurred in the preparatory zone, during which some of the fuel enters into the combustion chamber when the ignition is not initiated. The time duration between the start of injection to start of combustion is called as ignition delay period. The fuel properties influence the ignition delay and it is mainly governed by the cetane number of fuels.
SC
Comparisons between ignition delay with brake power for different geometries were shown in Figure 17 (a), (b), (c). The ignition delay of CIME blend is significantly lower than diesel
M AN U
and decreases gradually with increasing percentages of load. It is also noticed that the ignition delay decreases with increasing compression ratio. At low compression ratio, the amount of heat released is reduced which decreases the in-cylinder temperature, causes higher delay period. This reduction is caused mainly due to increase of cetane number and presence of oxygen in CIME blends [15]. This trend is in close agreement with diesel. The ignition delay ranges between 18°CA to 12°CA for diesel in HCC and subsequently
TE D
followed by 16°CA to 8°CA and 16°CA to 10°CA, for CIME in MHCC and HCC respectively. Besides, it is noticed that ignition delay decreases with load due to higher combustion temperature and dilution of exhaust gas at high loads [4].
EP
4. Emission Analysis
4.1 Carbon Monoxide (CO) emissions
AC C
The formation of CO and HC emissions are primarily due to the deficiency of oxygen in the fuel rich mixture during intermediate stages of combustion, which leads to incomplete combustion of the fuel [4]. The existence of local heterogeneity of the mixture could be one of the major dominant factors for the formation of CO. Figure 18 (a), (b), (c) illustrates the CO emission variations with respect to BP for different geometries. The trends of CO emission with different geometries are similar to diesel in conventional HCC geometry. The CO emissions at no load condition found to be higher for different geometries tried with B20 and diesel. This may happen due to short residence time and low ignition temperature during idling operation of the engine. However, with an increase in load, during combustion that completes the oxidation of CO to CO2 results increase in-cylinder gas temperature. In
ACCEPTED MANUSCRIPT comparison with diesel in conventional HCC, the raise in CO at full load condition is lower for HCC and MHCC fueled with B20. Among all, MHCC provides the lowest CO emissions. The lower value of CO emission could be due to the lesser heterogeneity of the mixture that minimizes the CO emissions and the quantity of air utilized with modified geometry is greater which enhances the combustion inturn reduces CO emission. Also, the presence of
RI PT
oxygen in CIME blend promotes the oxidation process during expansion that mitigates CO further. In addition, the formation of squish in this geometry, which dominates the higher heat loss can, minimizes the CO formation. In a similar report, Yang et al. [16], on evaluating the effects of piston bowl geometry on combustion and emission
SC
characteristics of kapok biodiesel fueled diesel engines stated that due to higher squish formation in a combustion geometry can minimizes the CO formation.
At higher
compression ratio 18, a reduction in CO, at part load, in MHCC is noticed as 0.02%, which is
M AN U
50% lower than in HCC emission while around 62% in the case of diesel in conventional bowl. At full load condition and at CR18, a reduction in CO is noticed as 35% and 50% for MHCC when compared with HCC and diesel in HCC. 0.14
HCC_DL
0.08 0.06 0.04 0.02 0 0.11
1
1.8
2.7
HC (ppm)
CO (%)
60
MHCC
0.1
TE D
0.12
HCC MHCC HCC_DL
70
HCC
CR16
CR16
50 40 30 20 10
0
3.5
0.11
1
0.14
CR17
CO (%)
0.12
2.7
3.5
Fig. 19 (a) Emissions: HC Vs BP at CR16
HCC
80
MHCC
70
HCC_DL
60 HC (ppm)
AC C
Fig. 18 (a) Emissions: CO Vs BP at CR16 0.16
1.8 BP (kW)
EP
BP (kW)
0.1 0.08 0.06
50 40 30
0.04
20
0.02
10
0
HCC MHCC HCC_DL
CR17
0 0.11
1
1.8
2.7
3.5
BP (kW)
Fig. 18 (b) Emissions: CO Vs BP at CR17
0.11
1
1.8
2.7
3.5
BP (kW)
Fig. 19 (b) Emissions: HC Vs BP at CR17
80
HCC CR18
MHCC
70
HCC_DL
60
HCC MHCC HCC_DL
CR18
50 40 30 20 10 0
0.11
1
1.8
2.7
3.5
BP (kW)
0.11
RI PT
0.2 0.18 0.16 0.14 0.12 0.1 0.08 0.06 0.04 0.02 0
HC (ppm)
CO (%)
ACCEPTED MANUSCRIPT
1
1.8
2.7
3.5
BP (kW)
Fig. 19 (c) Emissions: HC Vs BP at CR18
SC
Fig. 18 (c) Emissions: CO Vs BP at CR18
M AN U
4.2 Hydro Carbon (HC) emissions
Figure 19 (a), (b), (c) illustrates the HC emission variations with respect to BP for different geometries. It is observed that for the entire load range, HC emissions increased with increase in load for B20 but, it is lower than diesel. This may occur due to comparatively minimum availability of oxygen for the most amount of fuel injected at higher loads. This is reliable for higher compression ratio as well. At a lower compression ratio, low
TE D
temperature exists in the cylinder, due to poor combustion, which causes more HC emissions. This is in close agreement with the results obtained by Raheman and Ghadge [17] on their studies on diesel engine with variable compression ratio using biodiesel. Nevertheless, at a higher compression ratio, high temperature is developed and
EP
contributed towards better combustion that reduces the HC emissions. It is noticed that MHCC provides the lower HC emissions than HCC and diesel in HCC. This may attribute
AC C
due to better mixture formation and better air turbulence within the piston chamber. Similar results have obtained by Ghobadian et al. [18] in their studies on performance and emission parameters of a diesel engine using waste cooking oil biodiesel. A decrease of 10.14% and 13.8% has been remarked in the case of MHCC when compared with HCC and diesel in conventional piston respectively. 4.3 Oxides of Nitrogen (NOx) emissions It is clear that the formation of oxides of nitrogen is primarily due to nitrogen oxidation by oxygen in ambient air.
The oxygen rich combustion at elevated temperature and long
residence time are the causes for NOx formation. The NOx formation is totally controlled by
ACCEPTED MANUSCRIPT the amount of fuel burned and it is being greater at maximum load condition. During combustion the local temperature and the concentration of oxygen within the spray envelope is increased which may favor the increase in NOx emissions. Test results showing the variation of NOx emission with BP are plotted in Figures 20 (a), (b), (c), for a different piston bowl configuration at full load using B20. The NOx emission for the geometry MHCC is
RI PT
greater than HCC and diesel in HCC. This may be due to better fuel-air charge formation in the cylinder. However, due to lower ignition 1200
5
600 400
M AN U
1 0
0 0.11
1
1.8
2.7
0.11
3.5
BP (kW)
Fig. 20 (a) Emissions: NOx Vs BP at CR16 1400
1
1000 800 600 400 200 0 1
1.8
2.7
CR17
5
3.5
3 2 1 0 0.11
3.5
1
2.7
3.5
AC C
Fig. 21 (b) Emissions: CO2 Vs BP at CR17
1400
7
HCC MHCC HCC_DL
6
HCC MHCC HCC_DL
CR18
5 CO2 (%)
1000
1.8 BP (kW)
Fig. 20 (b) Emissions: NOx Vs BP at CR17
CR18
HCC MHCC HCC_DL
4
BP (kW)
Nox (ppm)
2.7
Fig. 21 (a) Emissions: CO2 Vs BP at CR16
6
EP
0.11
1200
1.8 BP (kW)
7
HCC MHCC HCC_DL
CR17
TE D
Nox (ppm)
3 2
200
1200
4
SC
CO2 (%)
800
HCC MHCC HCC_DL
CR16
6
CO2 (%)
NOx (ppm)
1000
7 HCC MHCC HCC_DL
CR16
800 600
4 3
400
2
200
1
0
0 0.11
1
1.8
2.7
3.5
BP (kW)
Fig. 20 (c) Emissions: NOx Vs BP at CR18
0.11
1
1.8
2.7
3.5
BP (kW)
Fig. 21 (c) Emissions: CO2 Vs BP at CR18
ACCEPTED MANUSCRIPT delay for biodiesel compared with diesel may be improved combustion occurring at bTDC. Due to complete combustion, the in-cylinder temperature increases, which, increase the NOx formations. The study carried out by Mohan T. Raj et al. [17] and Kirkude [18] have substantiated these findings for the cause of the increase in NOx formation. The NOx emissions noticed as 1186, 1198 ppm for HCC and MHCC respectively, while 1085ppm in
RI PT
the case of diesel in a conventional HCC bowl. For B20 fuel, the rise in NOx for HCC and MHCC are found to be 9.3% and 10.4% than diesel. 4.4 CO2 emissions
SC
The fluctuations of CO2 with respect to brake power with different geometries are shown in Figure 21 (a), (b), (c). It is found to be maximum CO2 emissions from the diesel fuel is 6.4% at higher compression ratio 18. However, it is noticed as 4.9%, 5.7% for B20 in MHCC, HCC
M AN U
respectively. It shows the predicted result in accordance with better mixture formation and
HCC MHCC HCC-DL
60 50
CR 16
TE D
Smoke opacity (%)
70
40 30 20 10 0 0.11
1
Smoke opacity (%)
spray qualities of the blend for the given compression ratio.
1.8
2.7
70 60
HCC
50 40
CR 17
MHCC
HCC-DL
30 20 10 0
3.5
0.11
1
EP
Brake Power (kW)
AC C
Fig. 22 (a) Emissions: Smoke Vs BP at CR16
Smoke opacity (%)
HCC
50
MHCC
40
HCC-DL
CR 18
30 20 10 0 0.11
2.7
3.5
Fig. 22 (b) Emissions: Smoke Vs BP at CR17
70 60
1.8
Brake Power (kW)
1
1.8
2.7
3.5
Brake Power (kW)
Fig. 22 (c) Emissions: Smoke Vs BP at CR18
ACCEPTED MANUSCRIPT 5. Conclusion The experimental investigations, carried out to investigate the effects of in-cylinder parameters using modified piston bowl geometry namely MHCC and compare it with a conventional piston bowl (HCC) using CIME-diesel blend (B20) and diesel fuel. Some of the
RI PT
significant conclusions from the results of the present investigations are: CIME blended with diesel (B20) results in better combustion and emissions associated with modified piston geometry (MHCC) than diesel. Increased CR results in substantial improvement in peak pressure with different geometry. At full load condition, higher peak
SC
pressures are noticed with B20 in MHCC than HCC and diesel in HCC. It shows that, MHCC provides 4.7% rise in peak pressure than diesel in HCC that may be due to shorter ignition delay period. The NHR values are decreased irrespective of the geometry with a rise in CR.
M AN U
The data chart interprets that the NHR values decrease with increase in compression ratio regardless of the geometry. There are 5.6% and 17% reduction in heat release rate was noticed for HCC and MHCC compared with diesel in HCC. A reduction of 35%, 13.8% and 34.1% emissions such as CO, HC and Smoke opacity respectively, are noticed at higher compression ratio 18 with MHCC geometry when compared to diesel. Oxides of
TE D
nitrogen are slightly higher for MHCC and HCC geometry fueled with B20 than diesel. This may attribute to the better combustion that enhances the turbulence of air-fuel mixture in momentous manner using modified piston. The modified piston geometry (MHCC) having lower bowl diameter to depth ratio makes higher squish flow, which enhances the mixture
EP
preparation during compression process. Overall, it can be seen that the output data from modified piston indicates a very significant improvement in almost all aspects than the
AC C
standard piston.
Acknowledgements
The authors would like to express their acknowledgement to All India Council for Technical Education (AICTE), India for providing the financial support to procure the equipment for this work, under MODROB scheme with Grant No.8024/RID/BOR/MOD-824/2009-10 for pursuing this research work at Sri Sairam Engineering College, Chennai.
ACCEPTED MANUSCRIPT
AC C
EP
TE D
SC
M AN U
CI Compression Ignition CIME Calophyllum Inophyllum Methyl Ester CR Compression Ratio SOC Start of Combustion EOC End of Combustion ID Ignition delay RPR Rate of Pressure Rise NHR Net Heat Release AV Acid Value VCR variable compression ratio °CA degree crank angle Q heat released during combustion mcri mass flow at the crevice Cv specific heat at constant volume CPmax maximum cylinder pressure U sensible heat energy sensible enthalpy of injected fuel hf Qw heat transferred from the cylinder Cp specific heat at constant pressure aTDC after Top Dead Centre bTDC before Top Dead Centre TDC Top Dead Centre BDC Bottom Dead Centre CO carbon monoxide CO2 carbon dioxide UBHC unburned hydro carbons NOx Oxides of nitrogen
RI PT
NOMENCLATURE
ACCEPTED MANUSCRIPT References [1] An, H., Chou, S.K., Li, J., Maghbouli, A., Yang, W.M., 2014, Effects of piston bowl geometry on combustion and emission characteristics of biodiesel fueled diesel engines, Fuel 120, 66–73. [2]
Annamalai, K., Jaichandar, S., 2012, Influences of re-entrant combustion chamber
RI PT
geometry on the performance of Pongamia biodiesel in a DI diesel engine, Energy 44, 633 640.
[3] Anand, T.N.C., Prasad, B.V.V.S.U., Ravikrishna, R.V., Sharma, C.S., 2011, High swirlinducing piston bowls in small diesel engines for emission reduction, Applied Energy
SC
88, 2355–2367.
[4] Annamalai, K., Jaichandar, S., 2012, Effects of open combustion chamber geometries on the performance of Pongamia biodiesel in a DI diesel engine, Fuel, 98, 272-279. Hadi Taghavifar, Samad Jafarmadar, Shahram Khalilarya, 2014, Engine structure
M AN U
[5]
modifications effect on the flow behavior, combustion, and performance characteristics of DI diesel engine, Energy Conversion and Management 85, 20–32. [6]
Feihu Wang, Kunpeng Ji, Shengli Wei, Xianyin Leng, Xin Liu, 2013, Numerical analysis on the effect of swirl ratios on swirl chamber combustion system of DI diesel
[7]
TE D
engines, Energy Conversion Management, 75, 184 – 190.
Lakshmi Narayana Rao, G., Mallikarjun, M.V., Venkata Ramesh Mamilla,
2013,
Effect of Combustion Chamber Design on a DI Diesel Engine Fuelled with Jatropha Methyl Esters Blends with Diesel, Procedia Engineering 64, 479 – 490. Nagarajan, G., Pasupathy Venkateswaran, S., 2010, Effects of the Re-Entrant Bowl
EP
[8]
Geometry on a DI Turbocharged Diesel Engine Performance and Emissions—A CFD
AC C
Approach, Journal of Engineering for Gas Turbines and Power, ASME, Vol. 132 / 122803-1. [9]
Das, L.M., Dilip Kumar Bora, Gajandra Babu, M.K., 2008, Performance of a mixed biodiesel fuelled diesel engine, Journal of scientific & Industrial research, 67, 73-76.
[10] Jayaraj, S., Muraleedharan, C., Ramadhas, A.S., 2006, Theoretical modelling & experimental studies on biodiesel-fuelled engine, Renewable energy, 31, 1813-1826. [11]
Venkanna, B.K., Venkataramana Reddy, C., 2011, Influence of Injector opening pressures on the performance, emission & combustion characteristics of DI diesel engine running on calophyllum inophyllumlinn oil (honne oil), International Journal of Renewable Energy, 6, 15-24.
ACCEPTED MANUSCRIPT [12] Atabani, A.E., Chong, W.T., Irfan Anjum Badruddin, keat Teong Lee, Masjuki, H.H., Mahila, T.M.I., 2013, Investigation of physical and chemical properties of potential edible and non-edible feedstock’s for biodiesel production, Renewable and Sustainable Energy Reviews 21, 749-755. [13] John B.Heywood, 1988, Internal Combustion Engine Fundamentals, 2/e, McGraw-Hill
RI PT
International Editions, Singapore, pp 385, 43. [14] Pundir, B.P., 2010, I.C.Engines Combustion and Emissions, Narosa Publishing House Pvt. Ltd. New Delhi.
[15] Altun, S., Oner, C., Biodiesel production from inedible animal tallow and an
Applied Energy 2009; 86:2114 – 2120.
SC
experimental investigation of it use as alternative fuel in a direct injection diesel engine,
[16] Lee, P.S., Saravanan, C.G., Vallinayagam, R., Vedharaj, S.,Yang, W.M., 2015,
M AN U
Optimization of combustion bowl geometry for the operation of kapok biodiesel – Diesel blends in a stationary diesel engine, Fuel 139, 561 – 567. [17] Ghadge, S.V., Raheman, H., 2008, Performance of diesel engine with biodiesel at varying compression ratio and ignition timing, Fuel 87, 2659–2666. [18] Ghobadian, B., Najafi, G., Nikbakht, A.M., Rahimi, H., Yusaf, T.F., 2009, Diesel
TE D
engine performance and exhaust emission analysis using waste cooking biodiesel fuel with an artificial neural network, Renewable Energy 34, 976–982. [17] Mohan T Raj, Murugumohan Kumar K Kandasamy, 2012, Tamanu oil - an alternative fuel for variable compression ratio engine, International Journal of Energy and
EP
Environmental Engineering 3 – 18.
[18] Atul S. Padalkar, Jagannath Hirkude, 2014, Experimental investigation of the effect of
AC C
compression ratio on performance and emissions of CI engine operated with waste fried oil methyl ester blend, Fuel Processing Technology 128, 367–375. [19] Agudelo JR., Lapuerta M, Rodriguez-Fernandez J, 2008, Diesel particulate emissions from used cooking oil biodiesel, Bioresource Technology 99, 731–40.
ACCEPTED MANUSCRIPT HIGHLIGHTS
RI PT
SC M AN U
•
TE D
•
EP
•
Combustion and emission characteristics of calophyllum inophyllum methyl esters were studied. The effects of different piston bowl geometries were investigated explicitly using biodiesel blend. Squish is more and enhanced charge preparation occurs in the modified piston bowl geometry. CO, HC and smoke emissions were significantly reduced when compared with conventional piston geometry.
AC C
•
