Availability analysis, performance, combustion and emission behavior of bael oil - diesel - diethyl ether blends in a variable compression ratio diesel engine

Renewable Energy 119 (2018) 235e252

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Availability analysis, performance, combustion and emission behavior of bael oil - diesel - diethyl ether blends in a variable compression ratio diesel engine M. Krishnamoorthi*, R. Malayalamurthi Department of Mechanical Engineering, Government College of Technology, Coimbatore, 641013, Tamil Nadu, India

a r t i c l e i n f o

a b s t r a c t

Article history: Available online 5 December 2017

The aim of the present work is to experimentally investigate the effect of injection pressure (IP) and injection timing (IT) on the performance, combustion, and emissions of a compression ignition (CI) engine with aegle marmelos oil (bael oil) blends. This work includes the exergy analysis of diesel engine to maximize the work availability. The tests were conducted on a constant speed direct injection diesel engine fueled with ternary blends of bael oil, diethyl ether (DEE) and neat diesel (D) at various engine loads. When the engine was operated with B2 blend (60%Dþ30%bael oilþ10%DEE), there was an increase in brake thermal efficiency of 3.5% accompanied by a declination in oxides of nitrogen emissions by 4.7% at full load with 250bar IP. The B2 blend showed lower hydrocarbon emission by 7% as compared to that of neat diesel at full engine load with fuel IT of 23 before top dead center. With Increase in engine load, augmentation exhaust gas and cooling water availabilities lead to amplification of exergy efficiency with increasing load. The exergy efficiency of B2 fuel has found as 62.17% of fuel input at 230bar IP with 100% load. From results, B2 fuel exhibits the best performance and combustion characteristics. © 2017 Elsevier Ltd. All rights reserved.

Keywords: Diesel Bael oil Diethyl ether Injection pressure Injection timing Exergy

1. Introduction The combustion of fossil fuels in internal combustion (IC) engines is one of the major sources of air pollution and fossil fuel depletion. Biofuels are one of the most effective solutions for a global warming reduction and fuel needs [1,2]. In developing country like India, the requirement for diesel fuel is around six times more than that of gasoline. Subsequently, the search for the alternative to neat diesel is becoming inevitable and most of the nations spend an enormous part of trade profit on acquiring crude petroleum from other nations [3,4]. Straight vegetable oils (SVO) obtained from plant seeds/harvests can be utilized directly in diesel engines without any engine modifications. Vegetable oils and alcohols that can be derived from biomass are imperative substitute fuels for use in diesel engines [5e7]. In general, kinematic viscosity of SVO does not meet the necessities ASTM D396 standard, which sets a limiting value of 2e3.6 centistokes (cS) at 40  C and is 10e15 times higher than that of diesel [8e10]. Micro-emulsification is one of the easiest and cost-effective processes used to blend two liquid

* Corresponding author. E-mail address: [email protected] (M. Krishnamoorthi). https://doi.org/10.1016/j.renene.2017.12.015 0960-1481/© 2017 Elsevier Ltd. All rights reserved.

fuels by using butanol, DEE, octanol or hexanol as a surfactant [11,12]. It is a general technique to decrease the viscosity of alternative fuel blends to the required level in order to facilitate fuel injection. Furthermore, certain solvents can be blended with vegetable oils to give a stable performance at lower temperatures [13]. The peak heat release showed by SVO is lower than that of the fossil diesel fuel and emits considerably lower oxides of nitrogen (NOx) emissions, higher carbon monoxide (CO) and hydrocarbon (HC) emissions [14e16]. Basic and simple ideas to overcome these problems are to blend the SVO with diesel fuel; however, test result shows high carbon deposits in the engine, piston ring sticking, injector coking, lubrication oil thickening as an after effect of dilution and polymerization [17,18]. The reduction in ignition delay and peak cylinder pressure, and an increase in combustion period are observed for diesel and SVO blended with oxygenated additives [19,20]. At the point when DEE composition increased over 24%, heavier smoke was observed [21]. The utilization of DEE has added a few constraints, such as lower lubricity, higher unpredictability and lower miscibility which may increase unburned HC emissions [22,23]. The first law of thermodynamics does not consider the quality of energy content of a system whereas the second law of thermodynamics diagnoses losses and provides solutions for enhancing

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engine performance and engine efficiency [24]. Exergy investigation contributes for planning more effective thermal system and to diminish inefficiencies in thermal systems [25]. Exergy losses to unburned species increased considerably at lower equivalence ratio [26,27]. Azoumah Y et al., investigates the exergy analysis combined with gas emission analysis for the compression ignition (CI) engine fueled with biofuels [25]. Advancing of injection timing reduces the NOx emission sharply by the lower peak combustion temperature. Then again, an increase in smoke opacity was observed as a result of retarded injection timing [28,29]. Further retardation of injection timing leads to augments HC emission in naturally aspirated diesel engines. The increase in injection pressure resulted in higher NOx and CO2, and decreased opacity emissions [30]. Aegle marmelos is a tree native to India which is generally growing wild in sub-Himalayan tracts, West Bengal also found in central and south India. The bael seeds were collected from the Eastern Ghats of Tamil Nadu, India and bael oil is extracted from the bael seed using the mechanical press. It is grown all over India, predominantly in sanctuary gardens because of its status as a sacred tree; likewise northern Malaya and Srilanka [29]. Aegle marmelos is a moderate growing, medium-sized tree; grow up to 12e15 m, which has a place with rutaceac family. The bael core seeds yielded 49% oil during extraction. The seed oil has 12.5% of 12-hydroxyoctadec-cis-enoic acids alongside other normal fatty acids [3]. It copes up with a wide range of soil conditions (pH range 5e10) and it is tolerant to water logging and has unusually wide temperature resilience (0  Ce50  C). The organic fruits are 5e7.5 cm in diameter, elongated pyriform in shape, with a dark or yellow skin. The seeds (fifty or more in a fruit) are implanted in a thick gummy pulp. The seeds were dried overnight at 55  C in an oven to eliminate excess moisture. The dried seeds were weighted and powered. The resulting bael seed oil is light yellow in color. The properties of bael oils are iodine value - 94 mg iodine/g; saponification value 205 mg/KOH; higher heating value (HHV) - 40.04 MJ/ kg; lower heating value (LHV) 33.27 MJ/kg [30]. In this paper, effects of IP and IT are taken as operating parameters and the diesel engines performance, combustion, exhaust emissions and exergy analysis has been investigated in the VCR engine with bael oil blends. The second objective is to find the appropriate input parameters to the CI engine for optimal output behaviors such as performance, emissions, and exergy efficiency for specified fuel blends and diesel. 2. Materials and methods 2.1. Fuel properties The flash point and fire point, density, and kinematic viscosity are determined for different fuel blends according to ASTM D-93, ASTM D-1298, and ASTM D-445 respectively. The DEE of 99% purity purchased from neighborhood business agent. Bael oil is mixed with diesel and DEE in a blender unit and stirred at 500 rpm for 20 min and left for 30 min to reach equilibrium at room temperature before the experimental test. Ternary blends of diesel fuel(D) e bael oil (SVO) - Diethyl ether (DEE) as percentages (vol. %) of 70% D 20%SVO-10%DEE (B1), 60%D-30%SVO-10%DEE (B2), 50%D-40% SVO-10%DEE (B3) were selected in the soluble area of ternary phase. The feasible blending ratio of vegetable oil to neat diesel is lies between 20 and 40%. The addition of fuel additives has improved the ignition performance and reduces the exhaust emissions. The phase separation of the fuel blends results that cavitations in the fuel line and injector nozzle if more than 15% of DEE to diesel. The erratic combustion takes place the DEE addition more than 24% in blends [56]. The erratic combustion takes place the DEE addition more than 24% in blends. The flash and fire points

are determined by Pensky Martin closed cup fire point apparatus. The kinematic viscosity is measured by the redwood viscometer. The properties of diesel, bael oil and DEE and its blends are given in Tables 1 and 2. The ignition performance of fuel is of crucial importance for CI engines as insufficient ignition quality can lead to higher emissions. The cetane number (CN) is widely used for indicating the ignition performance and is arrived with the following formulae according to the volumetric concentration of each constituent [3]: P (1) Cetane number CNH ¼ CNi Xi Pi (2) Calorific value CVH ¼ CVi Xi i

Where CNH, CVH are the equivalent cetane number, calorific value of the blended fuel, while CNi is the cetane number of each constituent, Xi is the percentage of constituents and CVi is the calorific value of each constituent.

2.2. Experimental setup and procedure To evaluate the performance and emissions of the test fuels and compare with diesel fuel, the experimental runs were conducted in a single cylinder direct injection variable compression ratio (VCR) test engine. The engine used for the test was a Kirloskar VCR and specifications are shown in Table 3. The fuel injection pressure (IP) is varied by 210bar, 230bar, and 250bar, and the fuel injection timing (IT) varied by 21, 23 and 25 before top dead center (bTDC) and the injector has three nozzle holes located near the combustion chamber center. The injection pressure was taken as 210bar, 230bar, and 250bar. When using lower IP (below 210bar) for vegetable oil blended fuels in compression ignition engines, more unburned hydrocarbon and smoke opacity were observed due to inefficient atomization [28]. If using higher IP (250bar and above) caused elevated unburned hydrocarbon and soot due to wall quench of fuel particle which can be attributed to higher momentum of the fuel particles that hit the combustion chamber wall and piston crown surfaces [29]. The engine was connected to an eddy current dynamometer and suitable arrangements were made to acquire all the controlling parameters. HC, CO2, NOx and CO emissions were measured with the aid of Exhaust gas analyzer AVL DI 444 model (Table 4). Smoke opacity is measured with the aid of Smoke meter, model AVL437C (Table 5). The piezoelectric transducer (Kistler make) is positioned within the cylinder head and is used for measuring in-cylinder pressure measurement. Piezoelectric pressure transducers are suitable for measuring highly dynamic, dynamic and quasi-static pressure curves or pulsations. The flush mounted technique is used to fasten the transducer on the cylinder head to avoid passage effects. Signals from pressure transducer are fed to charge amplifier. The signals from crank angle encoder and charge amplifier are acquired using data acquisition system (DAQ). The device is ideal for test, control and design applications including portable data logging, field Table 1 Properties of diesel, bael oil and DEE. Property

Diesel

Bael oil

DEE

Element structure Density (kg/m3) Viscosity (cS) Auto ignition point ( C) Cetane number (CN) Pour point ( C) Fire point ( C) Lower heating value (kJ/kg) Chemically correct A/F ratio

C16H34 830 2.7 200e400 50 20 52-96 42800 14.9

C18H36O2 896 24.3 <370 51.7 5 260 36300 12.4

C2H5OC2H5 713 0.23 160 >124 110 35 33900 11.1

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Table 2 Properties of blended fuels. S.No

Blend

Kinematic viscosity (cS)

Density (kg/m3) at 32  C

Calorific value (kJ/kg)

Auto ignition temperature ( C)

Cetane number

1 2 3

B1 B2 B3

6.81 7.99 10.11

831 849 876

41218 40476 39734

298 307 314

54.6 56.2 57.8

Table 3 Technical specifications of the test engine. Type No. of cylinders/No. of strokes Rated power Bore (mm)/Stroke (mm) Type of ignition Compression ratio Injection pressure Injection timing Speed Diameter/no. of nozzle hole Dynamometer Cylinder pressure sensor Crank angle encoder Data acquisition system Fuel flow measurement Load cell Air flow measurement

KIRLOSKAR, VCR multi fuel, vertical, water cooled, direct injection, naturally aspirated engine 01/04 3.5 kW/diesel mode, 4.5 kW/petrol mode 87.5/110 CI 12 to 18 210 bar 22 bTDC 1500 Rev/min 0.3 mm/3 Eddy current dynamometer; Water cooled; Model- TMEC10; RPM 1500e6000; Make - Technomech Pvt., Ltd. Make- Kistler; Piezo electric sensor; Model eM111A22; Resolution-0.1psi; Sensitivity e 1mV/psi. Make- Kistler; Model- 2614C11; Speed range - 0e12000 rpm; Crank angle signal - 720* 0.5 . USB-6210; 16AI; 4DI; 4DO USB- multifunction I/O device; Make- National instruments. Differential pressure transmitter; Make- Broiltech; Model- FCM. Sensortronics make; Model-60001. Make-Wika; Model- SL1.

Table 4 Technical specifications of gas analyzer AVL DI444. Measured quantity:

Measuring range:

Resolution:

Accuracy:

Carbon monoxide (CO)

0-10% vol

0.01% vol

Hydrocarbon (HC)

0-20000 ppm vol

2000:1 ppm vol, >2000:10 ppm vol

Oxides of nitrogen (NOx)

0-5000 ppm vol

1ppm vol

Carbon dioxide (CO2)

0-20% vol

0.1% vol

<0.6% vol: ± 0.03% vol 0.6% vol: ±5% vol <200 ppm vol: ±10 ppm vol 200 ppm vol: 5% <500 ppm vol: ±50 ppm vol 500 ppm vol: ±10 <10% vol: ±0.5% vol 10% vol: ±5% vol.

Table 5 Technical specifications of smoke meter AVL 437C.

Measuring Range Accuracy & repeatability Resolution

Opacity

Absorption

Rpm

Oil Temperature

0-100% ±1% of full Scale 0.1%

0e99.99 m-1 Better than ±0.1 m-1 0.01 m-1

400-6000 1/min ±10 ±1

0-150  C ±2  C ±1  C

monitoring, and in-vehicle data acquisition. Combustion analysis data is generally represented on the basis of degree ( ) of crank angle. The crank angle encoder provides angle - TDC relationship, necessary for the calculation of any crank angle based result related to a combustion cycle. The various temperatures are measured using K-type thermocouple fitted on respective positions. The water flow is adjusted to 70 and 250 L per hour for the calorimeter and engine cooling respectively according to the instructions given by engine supplier. The number of shims used to adjust the static injection timing by mounting it under the seat flange of the fuel pump. For the change of injection timing, the TDC position is marked on the flywheel. The repeated operations to be done to attain the exact injection timing by slowly rotating and stopping the flywheel rapidly. The shim is added and removed to vary the original injection timing to attain the required injection timing. The shims are inserted and removed by the nozzle spring to vary the injection pressure [31]. The pressure gauge is

connected to the fuel injection line to measure the pressure ranging from 100 to 400bar. In the beginning, the test engine operated for 20 min without any load and after stabilization, the experiments are conducted with constant speed and variable loads of (0 kg, 2 kg, 4 kg, 6 kg and 8 kg corresponding to 0%, 25%, 50%, 75% and 100% engine load) at steady environment air intake temperature. 3. Energy analysis The fuel incorporates energy on its chemical composition and releases heat energy during oxidation in the IC engine. According to this approach, we can calculate the cooling water availability from the coolant inlet and outlet temperature, and the exhaust gas availability from the exhaust gas thermocouple measurements. If the IC engine is considered as a control volume (surrounded by control surface) then the heat energy flows from and to the engine can be determined as follows, The steady flow first law analysis

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equation [3] for this control volume will be

Qs ¼ Pshaft þ Qw þ Qeg þ Qmiss

(1)

The terms of Eq. (1) are explained below The energy supplied by fuel (Qs ) to the diesel engine given by,

Qs ¼ mf  QLHV

(2)

Where mf is the mass flow rate of fuel supplied to the engine, QLHV is the lower calorific value of the fuel. The output power delivered (Pshaft ) by the engine or the shaft power is given by the equation,

Pshaft ¼ 2  p  NðrpsÞ  TðN:mÞ

(3)

Where T¼ (w  r), w is the applied load to the engine by the dynamometer and r is the effective arm radius. The amount of heat, which is carried away by the cooling water (Qw ) is given by,

Qw ¼ mw  Cpw  ðT2  T1 Þ

(4)

Where mw is the mass flow rate (kg/s) of coolant, Cpw is the specific heat of coolant, T2 is the coolant outlet temperature and T1 is the coolant inlet temperature. If the necessary heat to increase the temperature of the total mass (ma and mg are the mass of air and exhaust gas) with respect to outside atmosphere temperature (Ta ) to the exhaust gas temperature (Tg ) is known the value of heat loss through the exhaust gases can be calculated. The average specific heat (Cpg ) of the exhaust gas is considered as the specific heat of air at mean exhaust temperature [32].

Qeg

    ¼ ma þ mg  Cpg  Tg  Ta

(5)

The amount of heat transfer to the engine cylinder wall is calculated by [51],

  dQ w ¼ hc  A  Tg  Tw dq

(6)

Where Tg is the combustion gas temperature, hc is the convective heat transfer coefficient between cylinder wall and combustion gas, Tw is the cylinder wall temperature and A is the convection heat transfer area. Hohenberg's relationship was used for finding the instantaneous heat transfer [3]. The instantaneous heat transfer coefficient across the walls for the engine was predictable using Eq. (8),

hc ¼

 0:8 ! 130  p0:8 c  vp þ 1:4 V 0:06  Tg0:4

(7)

The unaccounted losses (Qun ) is computed by the energy balance equation and is given by,

  Qun ¼ Qs  Pshaft þ Qw þ Qeg

(8)

mechanical, thermal and chemical equilibrium with its environment [3]. In this section, the equations are given, which deals with the exergy balance to the IC engine and its subsystems in order to evaluate the various processes irreversibilities. The chemical equilibrium is achieved at the condition when no system component can react with the component in the environment. This means that in the dead state, all the components of the working medium were oxidized or reduced to N2, O2, CO2, and H2O. The total exergy of a system is equal to;

Ex ¼ Exch þ Extm ¼ E  P0 V  T0 S 

kk X

m0i mi

(9)

i¼1

Where i is the chemical potential of species at the ture dead state, and mi is the mass of species i [44]. The fuel availability for hydrocarbon-based fuels (CyHz), which is special interest to IC engines applications by;

  y 0:042 Ain ¼ QLHV 1:04224 þ 0:011925  z z

(10)

Shaft availability,

 dV dAw  ¼ Pcyl  P0  J=deg dq dq

(11)

Where dV is the rate of change of cylinder volume based on crank dq angle, Po indicates the ambient pressure, Pcyl is the instantaneous cylinder pressure that are calculated from energy analysis about the w represents indicated work transfer [43]. engine process and dA dq Availability input converted to cooling water availability (Acw ),

 T Acw ¼ Qw  mw  Cpw  T0  ln 2 T1

(12)

Availability transferred to exhaust gases (Aeg ),

     Ta Pa  Reg  ln Aeg ¼ Qeg meg  T0  Cpg  ln T5 Peg (13) Where, Reg is the gas constant for exhaust gas, Reg ¼ Ru =molecular weight, Ru - universal gas constant, Pa is the atmosphere pressure and Peg is the exhaust gas pressure. The destroyed availability (Ades ) is determined from the availability balance equation as,

i h  Ades ¼ Ain  Ashaft þ Acw þ Aeg

(14)

The ratio of total availability recovered from the system to the total availability input into the system is called exergy efficiency or second law efficiency (hII or 3) [3]. The recovered availability includes Ashaft , Acw and Aeg .

availability recovered availability input   Ades ε¼1 Ain

ε¼

(15)

4. Availability (exergy) analysis The exergy evaluation referred to diverse forms of control to have different levels of capability toward the positive work output. Availability of a system is defined as the highest work that can be obtained from a system through its reversible process to a state of

5. Results and discussion In this work, the effect of ternary blends of diesel, DEE and vegetable oils on VCR engine performance and emissions were

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investigated and exergy analysis was carried out. 5.1. Energy and exergy analysis For the second law analysis, the experimental assessment data were retrieved at this point. The complete experimental matrix was repeated for at least three times to record average experimental data for analysis purpose. Abusoglu et al., investigates the thermodynamic analysis of the diesel engine cogeneration system. The results concluded that exergy efficiency of the plant was obtained as 40.72% of input exergy and the remaining 59.65% of the exergy input was lost [26]. Junnian et al., experimentally determined and computed the combustion efficiency for different injection timings. The results concluded that 36.8% of the total fuel exergy was converted to net indicated work and about 16% of the fuel exergy transferred out of the system due to heat flows [33]. The second law analysis depicts the energy balance between the different terms such as availability input in fuel, availability recovered as beneficial power, availability transferred through exhaust gas and cooling water and destroyed availability due to friction, irreversibility, and exergy efficiency [42]. Jonathan et al., reported that the hydrogen fuelled ICEs approximately 41.37% of the chemical exergy was converted to work exergy and 27.3% of fuel exergy lost occur due to heat transfer [47]. The augmentation within the gross work output availabilities improved the corresponding exergy performance. 5.2. Input availability Fig. 1 illustrates the input availability (kJ/min) as a function of the engine load for different injection pressure and injection timing. As the load increases, the richer fuel-air mixture supplied to the engine leads to increase in combustion temperature. The input availability for blended fuels was higher for 210bar IP, because of the lower utilization of fuel in oxidization process. The LHV and volatility of the SVO fuel were lower in comparison with that of diesel fuels. At 75% engine loads, the B1 fuel had lower input

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availability of 514.9 kJ/min at 230bar IP which was 5.1 kJ/min lower than that of neat diesel. This indicates that the better utilization of fuel permitted a greater amount of the availability converted to work availability. The bael oil blended fuel operations need more amount of fuel to be supplied to the engine to produce higher output power. The maximum Ain has been observed in B3 fuel as 707.29 kJ/min at 100% engine load with 250bar IP which was 7.2% higher compared to that of neat diesel. When advancing IT, Ain increased due to long ignition delay and produce smooth combustion process. The Ain has been found as 649.88 kJ/min in 21 bTDC and 687.06 kJ/min in 23 bTDC for neat diesel at 100% engine load with standard IP. The maximum Ain has been observed in B3 fuel as 712.4 kJ/min at 25 bTDC while operating at 100% engine load. The long ignition delay allows more amount of fuel supply into combustion chamber due to insufficient ignition temperature for bael oil blend [30]. However, at the higher cylinder gas temperature, the input availability slightly increased due to the effect of charge dilution in the combustion chamber [47]. 5.3. Cooling water availability Fig. 2 depicts the cooling water availability on the basis of kJ/min for diesel and blended fuel operations. Generally increase in engine load shows enhancement in the combustion process, augmentation in the combustion temperature, peak pressure and reduction in combustion duration [29]. The cooling availability of B1 blend fuel operations was lower due to the intensive cylinder wall loss. Cooling availability was 2e3% higher for B3 blended fuel operations as compared to diesel. This can be attributed to the higher chemical energy input during blend fuel and long after burning zone. The minimum Acw have been observed in B1 fuel which was 3.6% and 4% lower compared to diesel operation corresponding to 75% and 100% engine load with 210bar IP. The maximum Acw for IT (23 bTDC) has been observed in B3 fuel as 69.7 kJ/min at 100% engine load which was 1.9 kJ/min lower than that of neat diesel. The long combustion duration leads to more heat energy transfer to the

Fig. 1. Input availability Vs load.

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Fig. 2. Cooling water availability Vs load.

Fig. 3. Exhaust gas availability Vs load.

cylinder wall. 5.4. Exhaust gas availability Fig. 3 presents the exhaust gas availability against the engine load for different test fuels. Basically, the exhaust gas availability increased with the effect of engine loads. The increasing of engine load produce higher combustion chamber temperature and some

of the thermal energy leaves along with exhaust gas. The highest exhaust gas availability loss was observed in B3 fuel as 183.2 kJ/min at 250bar IP while operating at 100% engine load. This can be explained by the fact that fraction of fuel oxidization taking place in controlled combustion and after burning zone resulted in more availability loss with exhaust gas. Because of these more availability losses via exhaust gas, the efficiencies of blended fuel operations are lower than neat diesel [45]. The lower Aeg has been observed for

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B2 fuel at 23 bTDC with 100% engine load was 158 kJ/min which was 1.7 kJ/min lower than that of neat diesel. The maximum Aeg observed for B3 fuel at 21 and 25 bTDC with 100% engine loads were 171.2 kJ/min and 171 kJ/min. In the case of 21 bTDC operations, the combustion process takes place rapidly with lower ignition delay and for 25 bTDC IT condition the combustion process takes place in slow and long progressive manner. 5.5. Destroyed availability Fig. 4 illustrates the destroyed availability (kJ/min) as a function of engine load for different injection pressure and injection timing. The amount of destroyed availability (as a percentage of fuel input) was decreased with increasing load [33]. The availability destruction was massive (as a percentage of fuel input) at no load conditions due to the zero shaft availability from the engine [28]. At low loads of 20%e40%, lower combustion temperature resulted in lower destroyed availability [33]. The minimum destroyed availability of 35.02% of fuel input has been found in B1 fuel at 100% load with 230bar IP. Diesel fuel operation showed that the minimum destroyed availability of 37.2% of fuel input was observed at 100% engine load with 250bar IP. The effect of advancing fuel injection decreases the Ades at 75% engine loads compared to that of retarded (21 bTDC) injection timing. The Ades slightly increased for 23 and 25 bTDC injection timing compared to 21 bTDC at 100% engine load. At lower engine loads of 20% and 40%, poor combustion behavior of SVO blended fuels causes lower cooling water and exhaust gas availabilities i.e., higher percentage of destroyed availability [25]. It has been acknowledged that most of the exergy destruction was generated during the combustion process and a higher combustion pressure was helpful in reduction of exergy destruction. 5.6. Exergy efficiency Fig. 5 shows the exergy efficiency for various test fuels with

241

loads at different IT and IP. For lower engine loads, the exergy efficiency was lower for all test fuels. The shaft availability becomes zero for no-load engine operations and the major input availability was shared by the destroyed availability [25]. For 75% engine load, the B1 fuel had a maximum exergy efficiency of 61.95% at 230bar IP. For 100% engine load, the B2 fuel had 2.2% higher exergy efficiency compared to diesel which was 62.05% for neat diesel at 230bar IP. The maximum exergy efficiency has been found as 63.17% for neat diesel at 75% engine load with 25 bTDC. The maximum of 63.9% exergy efficiency was found in B1 fuel at 23 bTDC with100% engine load. On account of utilizing diesel fuel, the minimum exergy efficiency was observed as 29.02% of fuel input at no load with 210bar IP. The injection timing plays an important role in exergy analysis and affecting the combustion irreversibility through its effect of gas temperature and pressure. This demonstrates that blend fuel engine operations cannot be ignored on the basis of their lower efficiency in a diesel engine which was actually designed for the standard diesel fuel [25]. 5.7. Brake specific fuel consumption (BSFC) The variations of BSFC at different IT and IP for various test fuel is shown in Fig. 6. The lower in-cylinder temperature has been observed in lower engine loads, which leads to deterioration of combustion process [52]. The increase in the cylinder pressure and temperature with increasing engine load causes a decreasing trend in the BSFC [46]. The BSFC was higher for bael oil blended fuels when the engine is working at 25% loads. At 75% engine load, the B1 fuel had 3.2% lower BSFC than neat diesel in 250bar IP. At 100% engine load, the BSFC of B1 fuel was found to be 1.3% lower and B2 fuel almost same with neat diesel in 230bar IP. Meanwhile, advancing of IT increases the mass of fuel injection to the cylinder due to long ignition delay period. The rising of BSFC for blended fuel operations was due to deterioration of the combustion caused by ignition delay and slow-burning velocity of vegetable oil molecules [36]. The heating values are also lower for blended fuels because of

Fig. 4. Destroyed availability Vs load.

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Fig. 5. Exergy efficiency Vs load.

Fig. 6. Brake specific fuel consumption.

lower calorific value of vegetable oil and DEE. The BSFC value slightly increased for neat diesel at 250bar IP as compared to 230bar IP. The BSFC values increased slightly at the higher combustion temperature due to the effect of charge dilution of fuel [47]. For 25 bTDC operation, the BSFC values increased compared to

23 bTDC for D and B1 fuel. This was due to excess fuel supply taken due to long ignition delay period [31]. The lower BSFC was observed for B1 fuel as compared to B2 and B3 in 100% engine load with 25 bTDC IT. The advancing of injection fuel actually reduced the cylinder temperature during compression stroke by absorbing

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latent heat from the air. The addition of DEE improved properties of vegetable oil such as kinematic viscosity, autoignition temperature, cetane number and fire point [34]. 5.8. Brake thermal efficiency (BTE) The BTE for different test fuels has been plotted against engine load as shown in Fig. 7. The increase of IP elevates BTE by better fuel spray formation in the combustion chamber. Increased BTE was observed at higher engine loads because of the higher combustion temperature and brake mean effective pressure. The higher BTE was observed in B2 fuel as 29.32% at 100% engine load with the IP of 250bar. This can be explained by the enhanced atomization, spray characteristics and air-fuel mixing, which brings about better combustion. From Fig. 7, it can be seen that B2 fuel has higher BTE compared to other test fuels with respect to 230bar and 250bar injection pressure. This indicates that the better combustion process happens with B2 fuels. Due to the decline in fuel consumption and the effective burning of HC molecules; the heat energy is obtained from the fuel at its maximum level. G.R Kannan et al., observed that the increasing injection pressure increases the BTE for vegetable oil blends, on account of the enhanced atomization which leads to easy oxidation of core of fuel droplets [35]. The BTE of diesel was higher than the bael oil blends for most of the engine loads at 21 bTDC IT because of lower viscosity and fire point of neat diesel. Advancing of injection timing leads to reduction for the rate of pressure rise because of the progressive combustion of fuel and air [50]. In this way, the BTE was influenced by the injection timing. The lower self-ignition temperature of the DEE blends helps in reduction of surface tension which produced complete combustion of the fuel and raising the BTE [12]. At 100% engine load, the BTE of B2 fuel was observed as 29.2% at the 230bar IP. The notable BTE observation at 100% engine load for the B1 blend was 28.6% at 21 bTDC IT. Similarly, the B1 fuel had higher BTE for the 21 and 23

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bTDC IT and 100% engine load. The addition of DEE in the bael oil and diesel reduced the peak cylinder pressure. This is because of the lower calorific value of the DEE and wider spray pattern. The increase in BTE for blended fuels (B1 and B2) at the advanced injection timing was due to the high heat release in the premixed combustion zone, and hence the peak pressure reaches closer to the TDC, which results in higher effective pressure to do more work [50].

5.9. HC emissions The variation of HC emission for the test fuels at different IP and IT are shown in Fig. 8. HC emission trend diminishes for all fuels with an increment of engine loads due to enhancement in combustion. The considerable reduction in HC emission was observed for B1 fuel compared neat diesel for 75% and 100% engine loads both in IT and IP conditions. At 100% load, B1 and B2 fuels had 16% and 11% lower HC emission than neat diesel at 250bar IP. At 75% load, B2 fuel had 10% lower HC emission compared to neat diesel at 230bar injection pressure. These reductions were due to complete burning of the blended fuel with an additive as to fulfill the combustion characteristics such as net heat release rate (HRR) and mass-burn fraction. The HC emissions increased for B3 fuels compared to other test fuels. Whenever the bael oil was blended with neat diesel, its viscosity increases which in turn to incomplete combustion and increased HC emission. The HC species oxidized due to higher cylinder gas temperature at the expansion stroke and higher heat release rate, which results in lower HC emission at higher loads. The impact of advancing injection timing increased the HC emission due to high fuel-air ratio, lower combustion temperature and size of fuel particles [29]. The maximum HC emission has been observed for B3 fuel as 18.9 ppm in 25% engine loads at 25 bTDC IT. The longer ignition delay permits more time for the fuel vapors to diffuse into the air and a higher percentage of

Fig. 7. Brake thermal efficiency.

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Fig. 8. Hydrocarbon.

fuel would be contained in the lean flame blowout region which causes hike in HC emission [30]. The B1 fuel showed lower HC emissions in 50% engine loads at 23 bTDC and 25 bTDC compared to other test fuels for the same operating conditions. The B1 and B2 fuel showed 4.2% and 7% lower HC emissions compared to diesel in 23 bTDC at 100% engine load. The vaporized fuel from the nozzle holes and sac volume enters the engine cylinder at low velocity later in the cycle during expansion stroke and has little time to mix with air when the combustion gas temperatures are still high. This part of the fuel remains generally unburned and is emitted in the exhaust [36]. 5.10. CO emissions Fig. 9 shows the influence of test fuels on the CO emissions for different values of IT and IP. Air-fuel proportion, injection pressure, engine load, type of fuel used and intake air temperature are the main operating parameters which influence CO emissions [11]. The primary factors for HC emissions were incomplete combustion of a fuel and insufficient oxygen for oxidation [8]. The increase in ignition delay and slow burning of bael oil at low engine loads affects the oxidation process [13]. At 250bar IP, the CO emissions were reduced because of better spray formation and air-fuel mixing. When the load is increased, the cylinder temperature also showed elevated trend; thus enhances the combustion process and leading to lower CO emissions. The CO emission has been lower for B2 fuel for all the engine loads compared with neat diesel at 250bar IP. At 75% load, the B2 fuel had 11.2% lower CO emission than the neat diesel in 250bar IP. At 100% engine load, B1 fuel showed 19.4% lower CO emission than that of neat diesel corresponding to 230bar IP. The reason for lower CO emission was enhancement of the combustion process with the effect of DEE [37]. Then again, the minimum CO emission of 0.02% happens in 23 bTDC IT at 75% engine load. At full engine loads, the CO emissions begin expanding again

and sharply rise, as still more fuel was injected to increase the power output. The retardation of IT gave a little higher CO emission than the original and advanced IT at 100% engine load. At advanced IT; the CO emission was lower due to longer combustion duration and increased oxidation between the carbon and oxygen molecules [50]. The advancing of 25 bTDC IT increased the CO emission for B3 fuels due to rich bael oil concentration and lower peak combustion temperature [30]. 5.11. NOx emissions The effect of test fuels on NOx emission for different IT and IP for as shown in Fig. 10. The high temperature burned gases due to the presence of high turbulence, rapidly mix with colder air or air-fuel mixture and the reduction reactions of NO2 back to NO results in higher concentrations of NOx [39]. Depending upon the burned gas temperature, the contribution of various reactions of the kinetics of NO formation also changes and henceforth the deviations might be seen from the simple stoichiometry adiabatic flame temperature [49]. There is a reduction scale observed in NOx emission when the engine operating with B1 fuel. At 75% load, NOx emission is 8.5% lower for B1 fuel compared to neat diesel in 210bar IP. Similarly at 100% load, B2 fuel had 4.7% lower NOx emissions with neat diesel in 230bar injection pressure. The maximum NOx emission has been observed for B3 fuel in 250bar IP as 581.2 ppm at 100% engine load which is 540.8 ppm for neat diesel. The formation of NOx starts in the burned gases produced from combustion close to stoichiometry and lean flammable mixtures during premixed combustion phase. The reduction in peak combustion temperature because of addition of the DEE which in turn reduces the calorific value of fuel blend helps in decreasing the NOx emissions [22]. The higher cetane number and reduced ignition delay period ensure the progressive burning; it helps in reducing the NOx emission [29]. It shows that NOx emission slightly decreased with DEE, because the combustion

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Fig. 9. Carbon monoxide.

Fig. 10. Oxides of nitrogen.

duration of the blend has been shortened. At 100% engine load, the minimum NOx emission observed in B1 fuel at 25 bTDC was 443.5 ppm which is 468.6 ppm for neat diesel. This advancing of IT

lowers the mass of the fuel accumulation before initiation of combustion process and reduces the initial combustion rates, hence decreasing the peak temperatures subsequently reducing the NOx

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formation [40,55]. At higher bael oil blend, NOx emission was increased due to the increasing combustion chamber temperatures at peak engine loads [24]. From Fig. 10, increasing proportion of bael oil in the fuel blends increased NOx compared to neat diesel with increasing of injection pressure [16]. Since the size of injected particles of vegetable oil was greater than that of diesel fuel if reduced the combustion efficiency [41].

combustion chamber. The dissociation process reduces the CO2 level by CO2 to CO conversion [54]. The engine operated with lower load condition, the CO2 emissions for diesel fuel was lowest compared with oxygenated B1 and B2 fuels [53]. At 75%e100% engine loads, significant improvement in oxidation process was observed due to higher combustion temperature. Throughout the engine operations, the CO2 emissions level increased considerably with the effect of IP and IT [35].

5.12. Carbon dioxide 5.13. Ignition delay period The complete combustion of fuel in the combustion chamber is directly indicated by the CO2 emission. Whenever hydrocarbon fuels are used in the combustion process, CO2 and water vapors were considered to be end products. From Fig. 11, the CO2 emissions initially lowered and subsequently increased with increase in engine loads for all the test fuels. It can be noted that the CO2 emission was lower for diesel fuel operation at 210bar IP and 21 bTDC. For 210bar IP operation, the fuel injection was not enough to provide a time for complete evaporation and subsequent combustion due to fuel particle size [47]. In the case of 230bar IP, the CO2 emission showed increasing trend for the test fuels. In 230bar IP, B2 fuel had 18% higher CO2 emission compares to neat diesel at 100% engine load. More CO2 is not much harmful to humans but it leads to ozone layer depletion and global warming. The advancing of IT increased the CO2 emissions due to long and progressive combustion. It also depends on the exhaust gas temperature [30]. The CO2 emission for biofuels combustion can be absorbed by the plants and kept a constant level in the atmosphere. The maximum CO2 emission has been found as 2.48% for B2 fuel at 25 bTDC with 100% engine load which was 8.81% higher than diesel fuel. The B1 fuel showed slightly lower CO2 emissions at 100% engine load with at 25 bTDC. This may be due to higher combustion temperature of about 1500  C, the dissociation of CO2 and H2O happened in the

Ignition delay period for test fuels at different IT and IP is shown in Fig. 12. The ignition delay period increased due to the lower compression gas temperature, higher enthalpy of evaporation and viscosity of fuel particles. Then again, the increase in IP decreases the ignition delay period by reducing the injected fuel droplet sizes and improved air-fuel mixing and combustion air temperature [36]. At lower engine loads the ignition delay was lower because of lower quantity of fuel injected and which forms a combustible mixture. In case of higher engine loads the amount of injected fuel is more, thus the fuel absorbs more heat energy and attain the self-ignition temperature little bit later during the compression stroke [7]. The results obtained by utilizing vegetable oils additionally showed a longer ignition delay with increasing number of double bonds [13]. The physical ignition delay has been shortened with increasing average number of double bonds. An average number of carbons do not seen to have a significantly effect on ignition delay. Since an increasing chain length can be expected to prompt a shorter chemical ignition delay, it is by all account neutralized by a longer physical ignition delay [39]. The bael oil blend has 1.1 CA longer ignition delay periods as compared to neat diesel. The maximum ignition delay has been observed as 13.6 CA in B3 fuel at 25 bTDC IT for mid and peak engine loads. In addition to the vegetable oils

Fig. 11. Carbon dioxide.

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Fig. 12. Ignition delay.

molecular structures, the initial combustion chamber pressure additionally affects the ignition delay. The addition of DEE to the fuel blends comes about the decline in ignition delay period and this is more prominent for all engine loads of engine operation [20]. The addition of DEE improves the cetane number and latent heat of vaporization of test fuels. At the higher loads, more amount of fuel is injected into the cylinder at higher pumping pressures tends to create better atomization. 5.14. Smoke opacity Fig. 13 shows the influence of test fuels on smoke opacity for different values of IT and IP. Smoke emissions can be decreased by accelerating the combustion process. The retardation of injection timing, the time available for the oxidation process decreased thereby increasing the smoke density. At 21 bTDC IP, the smoke is higher for B3 fuel compared to other injection timing. The advancing of injection timing increased combustion duration and permits more fuel oxidation of soot thereby decreasing smoke emissions [30]. The B1 fuel had 6.87% lower smoke opacity at 100% engine load with 23 bTDC. For 25 bTDC IT, the B1 fuel had 11.76% lower opacity than the neat diesel at 75% engine load. Further increment in the injection timing, there will be an increase in the ignition delay [47]. This might be on account of; the fuel injection takes place at a lower pressure and temperature in the engine cylinder. Thus, a substantial segment of injected fuel burns during the diffusion stage causing higher smoke density. Improved and complete combustion could be the reasons for obtaining lower smoke emission with the addition of oxygenated additives [41]. At 250bar IP, the smoke opacity for B1 fuel had 33.2% lower than neat diesel in full engine load. At 230bar IP, B2 fuel had 64.5% lower opacity compared to neat diesel in 25% engine load. Increasing the vegetable oil percentages in blended fuels also increased opacity emission at lower engine loads. The B3 fuel had higher opacity of

35.2 ppm at 100% engine load with 210bar IP. This was due to the higher percentages of vegetable oil in fuel blends that leads to incomplete combustion and carbon suspended particles [13]. 5.15. Cylinder pressure curve Fig. 14 indicates the cylinder pressure variation against crank angle (30 e40 connection with TDC) for different IP and IT. The rate of pressure rise was lower for reduced engine loads and hike in pressure raises was observed with increasing applied engine load. The maximum pressure release depends on the burned fuel fraction in the period of premixed burning phase [38]. The maximum cylinder gas pressure was observed for neat diesel operation at 100% engine load with 21 bTDC (210bar IP). The retardation of IT increases the peak pressure and reduces the combustion duration. The maximum pressure has been reduced by the addition of DEE to neat diesel. The shorter ignition delay and wider spray pattern have formed because of lower viscosity of the DEE in the blended fuel [20]. Optimum cylinder pressure rise was observed due to reduction in premixed fuel and lower LHV of the bael blends [20]. The peak pressure rate principally relies on the rate of combustion in the initial stage and fuel taking part in the uncontrolled combustion zone [45]. The effect of increasing the IP increased the rate of pressure rise by better combustion process and also too much increase in IP reduced the maximum pressure rise due to the effect of charge dilution in the engine cylinder. In case of blend B2, the combustion starts nearly the same crank position, which may reflect the pressure rise like that of neat diesel. For B3 fuel, the rate of pressure rise was lower compared to neat diesel and B1fuel due to the lower LHV and inefficient burning of fuel [22]. 5.16. Maximum cylinder gas temperature Average maximum cylinder temperature against crank angle

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Fig. 13. Smoke opacity.

Fig. 14. Cylinder gas pressure Vs crank angle.

(25 to 85 reference to TDC) is plotted in Fig. 15. At 25%e100% load, the maximum cylinder gas temperature was 2% lower for B1 fuel compared to neat diesel at 210bar IP with 21 bTDC. In 100% engine load, the average maximum cylinder temperature was lower for B1 fuel compared to other test fuels at 21 bTDC with 230bar IP. In 75% engine load, the maximum cylinder temperature was 1% higher for B2 fuel compared with neat diesel at 21 bTDC IT with 210bar IP. The increasing of IP increased the spray atomization and fuel droplet evaporation thereby reduced the ignition delay period. The progressive combustion takes place due to the lower ignition delay and thus peak temperature has been reduced. The reduced maximum combustion temperature was a reason for

reducing NOx emissions [41]. Low-temperature combustion is predicted to benefit overall engine efficiency, primarily because of reduced cylinder heat loss and potential of molecular properties of the expanding combustion gasses from dilute combustion phase to allow a greater amount of the energy to be extracted in the expansion stroke [40]. The retardation of IT increased the cylinder gas temperature by the fuel atomization which takes place at higher air temperature. The advancing of IT reduces the maximum cylinder gas temperature by progressive and long combustion period [50]. However, the thermal efficiency decreased little bit according to the Carnot theorem that is the efficiency based on the temperature limits [40].

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Fig. 15. Cylinder gas pressure Vs crank angle.

5.17. Average heat release rate The heat release rate was calculated by the first law of thermodynamics based on the data of the recorded cylinder pressure [47]. From the equation, the heat released by combustion of fuel is the sum of net heat release rate and heat absorbed by the cylinder wall. The heat release rate calculated by the following equation [48]:

    dQ g dV 1 dP dQ w P V þ þ dq g  1 dq dq g1 dq

(16)

where dQ is the amount of heat transfer rate (J/deg CA), g is ratio dq of specific heats, q is crank angle (deg), V is the instantaneous w is volume (m3), P is cylinder pressure (N/m2) of the cylinder, dQ dq wall heat transfer loss from Eq. (6) and Qw is the cylinder wall heat losses [3]. Fig. 16 shows the variations of average heat release rate (HRR) as a function of crank angle (CA) position (25 and 45 CA) for different IP and IT for test fuels. We observed the negative heat release for all fuels because of the combustible mixture express

cooling effect as a result of vaporization of the accumulated fuel during the ignition delay period in the compression stroke. As soon as the combustion phenomenon starts the HRR gets to be distinctly positive. The higher HRR for B1 fuel in beginning and peak combustion period compared to neat diesel at 21 bTDC with 21 bar IP. It can be observed that for each load considered, the average heat release curve for the B2 blend was little lower at the beginning than the corresponding one for diesel fuel. At 210bar IP, the ignition delay was high compared to 250bar IP. The longer ignition delay period allows more fuel accumulation before the initiation of ignition and produces peak cylinder due to more fuel burning [29]. After the ignition delay period, the premixed air/fuel mixture burns quickly after the diffusion combustion phase. At this point, the HRR was controlled with the aid of the charge of air/fuel mixing [37]. The rate of HRR has been higher for neat diesel operation due to the higher LHV compared to blended fuel. At lower loads, the rate of HRR was lower due to the smooth burning of a small quantity of fuel [47]. For higher engine loads, more amount of fuel was injected into the cylinder to produce higher output power by the rapid combustion process.

Fig. 16. Heat release rate Vs crank angle.

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Table 6 Uncertainties of some measured and calculated parameters. S.No

Parameter

Percentage uncertainties

1 2 3 4 5 6 7 8

NOx CO HC CO2 Smoke opacity Kinetic viscosity BTE BSFC

±0.1 ±0.01 ±0.1 ±0.3 ±0.5 ±1.3 ±1 ±1.5





6. Uncertainty analysis Uncertainties and errors in the experimental analysis may occur due to instruments selection, calibration, working condition, observation, environment and method of the tests [3,52]. The instruments for measurements are chosen with a view to keeping the experimental uncertainties as minimum as possible. The uncertainties have been calculated by Eq. (17);

sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi  2  2  2  Dx Dz Du Dw 2 ¼ þ …: þ þ þ…þ q x z u w

Dq

(17)

The instruments for measurements were chosen with a view to keeping the experimental uncertainties as minimum as possible. The probable errors in the stopwatch (Dt), speed indicator (Dn), graduated burette (DV), measuring scale (DR), and strain gauge type load cell (Dw) were 0.01sec, 5 rpm, 0.001 m and 0.1 kg respectively. Uncertainties of some measured and calculated parameters are given in Table 6. Based on the above values, the calculated engine performance is believed to be accurate within ±3% [35].

7. Conclusion Considering the need for alternative fuels, the experimental study was done in the present work in order to run the diesel engines with bael oil (SVO) and diesel with DEE additive. From the results of the comparative analysis, the diesel and vegetable oil blend with DEE ignition enhancer has the potential to reduce the harmful pollution and enhance the engine performance.  The maximum Ain has been observed in B3 fuel as 707.29 kJ/min at 100% engine load with 250bar IP which is 7.2% higher compared to neat diesel. The Ain has been found as 649.88 kJ/ min in 21 bTDC IT and 687.06 kJ/min in 23 bTDC IT for neat diesel at 100% engine load with standard IP.  The maximum Acw was observed in B3 fuel as 69.7 kJ/min at 100% engine load with 23 bTDC IT which is 67.8 kJ/min in neat diesel. The lower Aeg was observed in B2 fuel in 23 bTDC IT at 100% engine load was 158 kJ/min which is 159.7 kJ/min for neat diesel.  The minimum destroyed availability of 35.02% of fuel input has been found in B1 fuel at 100% load with 230bar IP. The Ades slightly increased for 23 and 25 bTDC injection timing compared to 21 bTDC IT at 100% engine load.  The maximum exergy efficiency has been observed as 63.17% for neat diesel at 75% engine load with 25 bTDC IT. The maximum of 63.9% exergy efficiency is found in B1 fuel corresponding to 23 bTDC IT and 100% engine load.  The higher BTE was observed in B2 fuel as 29.32% at 100% engine load with the injection pressure of 250bar. The notable BTE







observation at 100% engine load for the B1 fuel blend has 28.6% at the 21 bTDC injection timing. The considerable reduction in HC emission for B1 fuel compared neat diesel at 75% and 100% engine loads both in IT and IP conditions has been observed. The maximum HC emission has been observed in B3 fuel as 18.9 ppm at 25% engine loads with 25 bTDC IT. The B1 and B2 fuel have 4.2% and 7% lower HC emissions compared to diesel in 23 bTDC IT at 100% engine load. The CO emission has been lower for B2 fuel for all the engine loads compared with neat diesel at 250bar IP. Similarly, at 100% load, B2 fuel blend had 4.7% lower NOx emissions with neat diesel in 230bar injection pressure. The maximum NOx emission has been observed for B3 fuel in 250bar IP as 581.2 ppm at 100% engine load which is 540.8 ppm for neat diesel. The maximum CO2 has been found as 2.48% for B2 fuel at 25 bTDC IT with 100% engine load which was 8.81% higher than diesel fuel. At 75% and 100% engine loads, significant improvement in oxidation process was observed, which was due to higher combustion temperature. The maximum ignition delay has been observed as 13.6CA in B3 fuel at 25 bTDC IT for middle and peak engine loads. The B1 fuel shared 6.87% lower smoke opacity at 100% engine load with 23 bTDC IT. For 25 bTDC IT, the B1 fuel had 11.76% lower opacity than of neat diesel at 75% engine load. At 230bar IP, B2 fuel had 64.5% lower opacity compared to neat diesel at 25% engine load. The maximum cylinder gas temperature was observed for neat diesel operation at 100% engine load with 21 bTDC IT (210bar IP). At 100% engine load, the average maximum cylinder temperature was observed for B1 fuel compared with other test fuels at 21 bTDC IT with 230bar IP.

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Nomenclature ASTM: American society for testing and materials B1: 70% dieselþ 20% bael oilþ 10% DEE B2: 60% dieselþ 30% bael oilþ 10% DEE B3: 50% dieselþ 40% bael oilþ 10% DEE BSFC: brake specific fuel consumption BTE: brake thermal efficiency CA: crank angle CE: chemical energy CI: compression ignition CN: cetane number CO: carbon monoxide CO2: carbon dioxide CR: compression ratio D: neat diesel DEE: diethyl ether HC: Hydrocarbon HHV: higher heating value

HRR: heat release rate IC: internal combustion IP: injection pressure IT: injection timing L: exergy loss LHV: lower heating value NOx: oxides of nitrogen rps: revolution per seconds SVO: straight vegetable oil TDC: top dead center VCR: variable compression ratio w: engine load in Newton W: Watts 230bar IP: fuel injection pressure - 230bar 21 : IT e 21 before TDC Greek letter

g: ratio of specific heats hII or 3 : exergy efficiency q: crank angle Р: density y: viscosity (centistokes) Subscripts 0: dead state condition a: atmosphere condition cw: cooling water des: Destroyed eg: exhaust gas in: input