Experimental investigation on performance, emission behavior and exergy analysis of a variable compression ratio engine fueled with diesel - aegle marmelos oil - diethyl ether blends

Accepted Manuscript Experimental investigation on performance, emission behavior and exergy analysis of a variable compression ratio engine fueled with diesel- aegle marmelos oil - diethyl ether blends M. Krishnamoorthi, R. Malayalamurthi PII:

S0360-5442(17)30606-0

DOI:

10.1016/j.energy.2017.04.038

Reference:

EGY 10678

To appear in:

Energy

Received Date: 26 December 2016 Revised Date:

28 February 2017

Accepted Date: 9 April 2017

Please cite this article as: Krishnamoorthi M, Malayalamurthi R, Experimental investigation on performance, emission behavior and exergy analysis of a variable compression ratio engine fueled with diesel- aegle marmelos oil - diethyl ether blends, Energy (2017), doi: 10.1016/j.energy.2017.04.038. 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.

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Experimental investigation on performance, emission behavior and

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exergy analysis of a variable compression ratio engine fueled with

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diesel- aegle marmelos oil - diethyl ether blends

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M.Krishnamoorthia*, R.Malayalamurthib

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a

Department of Mechanical Engineering, Government College of Technology, Coimbatore 641013, India.

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b

Department of Mechanical Engineering, Faculty of Engineering, Government College of Technology, Coimbatore

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641013, India.

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Corresponding author E.mail:[email protected]

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Mobile: +91 9940772158

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Abstract

The intention of the prevailing effort is in the direction of experimentally look for the

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combined outcome of compression ratio and a number of nozzle holes on performance and

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emissions of a compression ignition engine by means of an emulsion fuel obtained from aegle

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marmelos (Bael) oil. This exertion consists of the exergy examination of compression ignition

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engine towards maximizing the work availability and decreasing the destroyed availability.

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Ternary blends of diesel - aegle marmelos – diethyl ether (DEE) within the proportion as

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percentages 100:0:0 (D), 70:20:10 (B1), 60:30:10 (B2), 50:40:10 (B3) became tested in a

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variable compression ratio (VCR) engine. When operating the diesel engine with B2, Brake

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thermal efficiency (BTE) of the engine is better by 4.3%, nitric oxides (NOx) emission has been

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reduced 3.9% at 100% load in compression ratio (CR) 17.5 with number of nozzle hole (NH) 5.

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The exergy efficiency of B2 fuel has been found 63.88% of fuel input at CR17.5 with 100%

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engine load. Increasing the number of nozzle holes improves the performance of the diesel

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engine fuelled with bael blends in terms of reduced brake specific energy consumption (BSEC),

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increased BTE and reduced emissions like hydrocarbon (HC), carbon monoxide (CO) and

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smoke.

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Keywords

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Diesel; Bael oil; Diethyl ether; Compression ratio; Nozzle hole; Exergy.

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1. Introduction

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The use of new, substitute and clean-burning fuels seeing that essential energy asset into

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internal combustion (IC) engines is a worldwide attention to accomplish lesser pollution and oil

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economy [1]. Biofuels are one of the renewable fuels, which is generally created from edible

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crops (groundnut, cottonseed, palm nut, etc.) and non-edible crops (calophullum inophullum,

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jatropha curcas, algae, bael, etc.)[2]. A long-standing use of hydrocarbon-based fossil fuels tend

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to increase unavailability nearby makes an alarm to discover environmentally suitable alternative

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energy sources, with easy accessibility and economic viability [3,4]. From plant seeds/harvests

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we got vegetable oils and can be utilized significantly in diesel engines. The diesel engines can

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be operated with substitute fuels such as biomass and it is imperative to use in engines which can

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be conveyed from vegetable oils and alcohols [5]. The purposes of interest on the grounds that

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renewable energy sources obtained from alcohols and vegetable oils are that may have lower

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exhaust pollutions because of containing the oxygen molecules in their synthetic structure [6].

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Aside commencing the renewability, the vegetable oils have a prominent flash point and

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lubricity, a lesser amount of sulfur and aromatic values, the superior biodegradability and non-

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poisonous compared to fossil fuels [7]. The kinematic viscosity of vegetable oils does not meet

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the diesel fuel standards as ASTM, which is 2-3.6 Centistokes at 40oC but the vegetable oils have

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10 -15 times of fossil diesel viscosity and it is reduced by the microemulsion process [8].

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Microemulsion method is used to combine the vegetable oils, diesel and alcohols such as diethyl 2

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ether (DEE), that have promising potential as biofuels for a CI engine. This is a common

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procedure to be employed the substitute fuel to accomplish their viscosity for diesel engine

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injection system [9]. Furthermore, the solvents can be blended with the vegetable oils and

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provide stable alternative fuels in minimum temperatures limits [10]. The peak heat release

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showed by straight vegetable oil (SVO) is lower than that of the fossil diesel and emits

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considerably lower NOx emissions, elevated CO and HC pollutions [11]. Attempts to make use

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of SVO (100%) as option fuels to diesel engine have experienced a fewer number of operational

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problems [12]. Fundamental and straightforward attempts to defeat these troubles by mixing the

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SVO with fossil diesel fuel; nevertheless, tests outcome said high carbon deposits in the engine,

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sticking in piston rings, injector coking, thickening of lubrication oil as an after-effect of

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intensity and polymerization [13].

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DEE is lacking in originality from ethanol as it is referred to as renewable energy and it

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becomes utilized in cold start assist during earlier occasion period of its utilization as a fuel and

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practically minimum proportion for the reason that of the lower energy content [14,15]. The

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engine will become insecure and heavier smoke discovered at the point when the DEE fraction

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increased further than 24% [16-18]. The first law of thermodynamics considers the quantity of

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energy content in a system. Exergy investigation contributes for planning added an efficient

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thermal system of numerous types and guiding efforts to diminish inefficiencies in thermal

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structures [19]. The exergy (second law of thermodynamics) diagnoses the sufferers and

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provides solutions for enhancing the engine performance and engine competence [20]. Exergy

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losses (L) to thermal loss decreases with lower theoretical air/fuel ratios and will increase just

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about linearly with advancing in combustion timing, and exergy losses to unburned species

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decrease noticeably at higher equivalence proportion [21,22]. Y.Azoumah et al., investigates the

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exergy analysis combined with gas emission analysis for the CI engine fueled with biofuels [19].

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In this paper, combined effects of compression ratio (CR) and a number of nozzle holes

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(NH) were taken as operating parameters and the diesel engine performance, combustion,

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exhaust emissions and exergy analysis were measured for VCR engine fuelling biofuels blends.

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The second objective is to find the appropriate input parameters (CR and NH) to the CI engine

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for optimal output behaviors such as performance, emissions and exergy efficiency for specified

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fuel blends and diesel.

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2. Materials and methods

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Aegle marmelos is a tree native of India which is growing wild inside the sub-Himalayan

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forest, west Bengal, middle India and Tamil Nadu. Aegle marmelos tree is cultivated all over

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India, predominantly within sanctuary gardens due to the position as a sanctified tree; likewise

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northern Malaya and Srilanka [23]. Aegle marmelos be a moderate growing, medium-sized tree;

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grow as much as twelve to fifteen meters, which has a place with rutaceac family. The bael core

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seeds containing almost fifty percent oil content. The bael oil has 12.5 percentages of 12

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hydroxyoctadec-cis-enoic acids together with normal fatty acids. It copes with a wide range of

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soil conditions (pH range 5-10) and is tolerant to water logging and has unusually wide

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temperature resilience (0°C to 50°C). The organic fruits are five to seven centimeter in diameter,

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stretched out pyriform in shape, with a darkish or yellow casing. The seeds of fifty or more in a

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fruit are implanted in a thick gummy mash. An oven used to dry the bael seeds at 55oC in a

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single day to take away excess moistures. After that, the dried seeds are weighted and powered.

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The mechanical extractor is used toward extract the bael oil and filtered with micron intensity.

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The resulting bael seed oil is light yellow in color. The properties of bael oils are iodine value-

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94mg iodine/g (it belongs to monounsaturated vegetable oil); saponification value -

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0.205g/KOH; higher heating value (HHV) - 40040kJ/kg; lower heating value (LHV) -

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36300kJ/kg [24,25]. The DEE of 99% purity purchased from neighborhood business enterprise

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agent. Bael oil became mixed with diesel and DEE fuel in a blender unit and stirred in an

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electromagnetic agitator at 500 rpm for 20 minutes and left for 30 minutes to accomplish

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equilibrium at room temperature before the experimental trial.

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2.1 Fuel properties

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First of all, the fuel properties of different fuel blends inclusive of flash and fire point,

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kinematic viscosity, and density are found out in line with ASTM D-93, ASTM D-445, and

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ASTM D-1298 respectively. The flash and fire points are determined by the use of closed cup

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fire point device. The redwood viscometer has been used to determine the kinematic viscosity of

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liquid fuels. The properties of diesel, bael oil, DEE and its blends are specified in Table 1 and

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Table 2. The crucial importance of a CI engines are detonation behavior of engines and

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unsatisfactory ignition quality of fuel may possibly lead to higher exhaust pollutions. The cetane

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number (CN) is extensively used to indicate the ignition performance and is arrived with the

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subsequent formulae based on the volumetric concentration of each component [7,51];

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(1) Cetane number CNH=∑ CN X

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(2) Calorific value CVH =∑ CV X

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Where CNH, CVH are the equivalent cetane number, calorific value of the blended fuel,

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while CNi is the cetane number of each constituent, Xi is the percentage of constituents and CVi

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is the calorific value of each constituent.

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2.2 Experimental setup and procedure

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To assess the overall performance of the test fuels as well as to compare with diesel fuel,

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assessments were performed in a single cylinder direct injection variable compression ratio test

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engine. The engine used for the test was a Kirloskar VCR and technical specifications are given

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in Table 3. The cylinder CR are varied by CR14, CR15, CR16, CR17.5 and number of fuel

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injector nozzle holes are varied by one, three, five [26,27]. The nozzle hole diameters are varied

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as a function of a number of nozzle holes (Table 3). The start of injection is same for all test

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fuels. When using lower compression ratio (below CR14) for vegetable oil blended fuels in

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compression ignition engines, emits more unburned hydrocarbon and smoke opacity [28]. The

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engine efficiency (diesel cycle) depends upon the compression ratio of the engine [19]. If using

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higher compression ratio (about CR18 and above) it could cause for higher oxides of nitrogen

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due to a higher temperature in the combustion chamber, unburned hydrocarbon and soot due to

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wall quench of fuel particle [34]. This happens due to lower clearance volume of the combustion

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chamber may cause for direct hitting of fuel droplets on the cylinder wall and piston crown

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[32,33]. Eddy current dynamometers with appropriate arrangements are linked to the engine in

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order to gather all controlling parameters. HC, CO2, NOx and CO emissions were measured by

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the aid of Exhaust gas analyzer AVL DI 444 model (Table 4). Smoke opacity is measured with

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the aid of Smoke meter, model AVL437C (Table 5). The piezoelectric transducer (Kistler make)

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is positioned within the cylinder head through a water-cooling arrangement and it is used for the

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inside cylinder pressure measurement. Piezoelectric pressure transducers are suitable for

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measuring highly dynamic, dynamic and quasi-static pressure curves or pulsations. The flush

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mounted technique is used to fasten the transducer on the cylinder head to avoid the passage

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affects. Signals from pressure transducer are fed to charge amplifier. The signals from crank

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angle encoder and charge amplifier are acquired using data acquisition system (DAQ). The

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device is ideal for test, control and design applications including portable data logging, field

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monitoring and in-vehicle data acquisition. Combustion analysis data is generally represented on

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the basis of degree (°) of crank angle. The crank angle encoder provides an angle and a TDC

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relationship, necessary for the calculation of any crank angle based result related to a combustion

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cycle. The various temperatures are measured using K-type thermocouple fitted on respective

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position. The water flow was adjusted to 70 and 250 liters per hour for the calorimeter and

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engine cooling respectively according to the instructions given by engine supplier. In the

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beginning, the test engine operated for 20 minutes without any load and after stabilization, the

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experiments were conducted with constant speed and variable loads of (0kg, 2kg, 4kg, 6kg and

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8kg corresponding to 0%, 25%, 50%, 75% and 100% engine load) at steady environment air

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intake temperature.

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3. First law analysis

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The fuel incorporates energy on its chemical composition and oxidized with oxygen or

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air in IC engine. Consequently, the fuel power converted into thermal energy and the amount of

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given input chemical energy (CE) is then converted predominantly into beneficial brake power

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or energy at crankshaft (Pshaft), thermal energy transformed into the cooling water (Qcw), thermal

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energy transformed to the combustion products (Qeg), Unaccounted energy loss (Qun) due to

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running auxiliary equipment, friction, radiation, combustion irreversibility, heat transfer to

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environment etc. In this analysis the amount of this energy acknowledged and evaluated on the

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idea of the first law of thermodynamics is described as below [30].

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The capacity of input energy (Qin) to the diesel engine is the energy content of its

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supplied fuel given by, •

Qin = mf × QLHV Watt

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where, mf is the mass flow rate of fuel supplied to the engine, QLHV is the calorific value of the

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fuel.

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equation,

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•

Pshaft = 2 ×π×N(rps/s)×w(N)×r(m) Watt

where, w is the applied load to the engine by the dynamometer, r is the effective arm radius. The amount of heat loss from the engine block to the cooling water (Qcw) is given by,

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Qcw = mw×Cpw × (T2-T1) Watt

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The energy converted to shaft power or the output power delivered (Pshaft) is given by the

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where, mw is the mass flow rate (kg/s) of coolant, Cpw is the specific heat of coolant, T2 is the

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coolant outlet temperature and T1 is the coolant inlet temperature.

The exhaust gas has thermal energy and it is transferred to the atmosphere is evaluated

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by, •

Qeg = meg×Cpg× (Tg-Ta) Watt

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where, meg is the mass of exhaust gas (mass flow rate of fuel burned + mass flow rate of air to

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the engine), specific heat capacity of the exhaust gas is mentioned by Cpg, Tg is the temperature

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of combustion gas at the exit, Ta is the outside ambient air temperature.

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In some case, the exhaust gas calorimeter used to measure the thermal energy from the

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exhaust side,

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Qeg=

m c ×  ×4 −3×5 −  5 −6 

Watt

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wherein mc is the mass flow rate of water into exhaust gas calorimeter, the calorimeter water

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inlet temperature is T3, T4 is the calorimeter water outlet temperature, the exhaust gas

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temperature at calorimeter inlet is T5, T6 is the exhaust gas temperature at calorimeter outlet and

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the atmosphere temperature is T0 (32°C) [19]. 8

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The unaccounted losses (Qun) is calculated by equating the energy balance equation and

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is given by,

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Qun = Qin – (Pshaft+ Qcw + Qeg) Watt

4. Availability (exergy or second law) analysis

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The perceptive of how the lack of energy occurs that will facilitate the findings to

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minimize the identical losses to improve the function of the engine in terms of power output and

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typical performance. The exergy evaluation referred to diverse forms of control to have different

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levels of capability toward the positive useful work output. This ability to carry out positive

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motorized work has been defined as availability. The availability of a system is defined as the

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maximum quantity of beneficial work accessible as soon as the system has brought into thermal,

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chemical and mechanical equilibrium through its environment by the way of reversible processes

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at the same time exchanging thermal energy with the surroundings only. In the IC engine,

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chemical availability of fuel containing the availability input (Ain) and it is transformed into

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further availability forms.

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In an engine, the input fuel availability is converted into the various forms [30]:

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Element availability of oil or input availability,

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where, 1.0338 is a constant for liquid diesel or blended fuels and varied for gaseous fuels [32]. Shaft availability,

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Ashaft/ dθ = p − p  × dv⁄dθ Watt

Availability input converted to cooling water availability (Acw),

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Ain = [1.0338×mf × QLHV] Watt

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Acw = Qcw − m × C! × T × ln 2' Watt 1

Availability transferred to exhaust gasses (Aeg), 9

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,

Aeg = Qeg+ (m)* × T × C!* × ln + - / – 1R )* × ln 1

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3-

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667 Watt

where, Reg is the gas constant for exhaust gas, Reg = Ru/molecular weight, Ru - universal gas

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constant, Pa is the atmosphere pressure, Peg is the exhaust gas pressure.

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The uncounted destroyed availability (Ades ) is determined from the availability balance

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equation as,

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Ades = [Ain – (Ashaft + Acw + Aeg )] Watt

Exergy efficiencies are beneficial in support of distinguishing the utilization of energy

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assets to facilitate thermodynamically usefulness from those that are less so. Exergy

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effectiveness moreover may be used to find the inefficiency of engineering device and taken

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necessary accomplishment to expand the performance of a thermal system. The ratio of total

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availability recovered from the system to the total availability input into the system is called

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exergy efficiency or second law efficiency (ɳ99 or ɛ) [31]. The recovered availability includes

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Ashaft, Acw and Aeg. :;:<:=<>? @)AB;)@)C

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ɳ99 =

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ɳ99 = 1–

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5. Results and discussion

In this work, the final results of ternary blends of vegetable oil, DEE and diesel oils on

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VCR engine overall performance, emissions are investigated and exergy analysis is done. The

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characteristics of the engine with ternary blends have been depicted as BTE, BSEC, NOx, CO,

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CO2, HC and smoke opacity.

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5.1 Energy and Exergy analysis

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For the second law analysis, the experimental assessment data are retrieved at this point.

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The complete experimental matrix was repeated for at least three times to record average

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experimental data for analysis purpose [32]. The availability balance for both the diesel and three

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blend fuel operations as a function of the percentage of engine load is developed. The second law

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analysis depicts the energy balance between the different terms such as availability input in fuel,

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availability recovered as beneficial power, availability transferred through exhaust gas and

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cooling water, and destroyed availability due to friction, irreversibility, and exergy efficiency

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[30]. The shaft availability is almost equal to the engine output thermal efficiency [32]. The

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exhaust gas availability and cooling availability are increased because of the progressed

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combustion characteristics by the addition of DEE, particularly in higher engine loads.

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Furthermore, as increasing engine loads accelerate the shaft availability because it is a function

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of power output [19]. The availability outcome confirmed that the bael oil blended fuel

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operations created more increments inside the collective exhaust gas and cooling water

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availabilities when the load is increased. The increase within the gross work output availabilities

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improved the corresponding exergy performance [20].

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Fig.1-3 indicates the input availability (kJ/min) and cooling water availability (kJ/min) as

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a function of the percentage of engine load for different compression ratio and nozzle holes. The

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wealthy fuel/air combination build up in the combustion chamber it creates combustion

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temperature rise while increasing engine load [27]. Since mentioned previously LHV and

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volatility of the SVO fuel are lower in comparison with that fossil diesel. The bael oil blended

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fuel operations the more amount of fuel is supplied to the engine to generate higher output power

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[25]. This permitted a greater amount of the energy conversion to work availability accessible

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from engine [28]. Thus the way more fuel is necessary for required power development in the

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engine. The input availability has been decreasing with the increasing of compression ratio up to

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CR16. From Fig.1, the availability input (Ain) for NH1 has a maximum of 673.4kJ/min for B3

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fuel observed in CR17.5 with 100% engine load. This is due to more amounts of fuel supplied to

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the engine during blended fuel operations [10]. This is because increasing the CR in the engine

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increases the cylinder temperature and enhances the combustion process [28]. However, at the

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higher compression ratio, the input availability slightly increased due to the effect of charge

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dilution in the combustion chamber due to higher temperature [34]. The maximum input

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availability for diesel mode was 671.7kJ/min at CR17.5 with 100% engine load. The minimum

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input availability of 653.2 has been observed at B1 blend at CR17.5 with 100% engine load. This

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is due to decreasing the viscosity of fuel by the adding of DEE and efficient burning of fuel [15].

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Fig.2 indicates the input availability for three nozzle hole injector. The input availability for NH3

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was slightly lower compared to single hole nozzle operation. This is due to better atomization of

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fuel into the combustion chamber [33]. Fig.3 shows the input availability for five hole nozzle

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operation. The maximum input availability has been found in B1 fuel at 100% engine load with

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CR17.5 was 677.8kJ/min and the minimum of 659.8kJ./min for neat diesel at CR16 with 100%

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engine load.

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The cooling water availability (Acw) is increased with the effect of increasing the CR and

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load. Generally, an increase in the compression ratio results in enhancement of combustion

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process, increasing the combustion temperature and peak pressure, and reducing the combustion

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duration [26]. The increasing of a number of nozzle holes increases the atomization of fuel and

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air/fuel mixing [34]. The cooling availability for the blended fuel operation is lower due to the

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intensive cylinder wall loss [35]. Cooling availability in the range 8-12% of fuel input is

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accessible in blended fuel operations compared to about 7-10% to that of diesel [29]. From Fig.1,

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the cooling water availability was lower for lower CR operations and increased with the

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increasing of compression ratio for all test fuels. This is due to increasing of heat transfer to the

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cylinder wall [30]. From Fig. 2, the maximum cooling water availability loss has been observed

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as 60.7kJ/min for neat diesel at CR17.5 with 100% engine load. This directly indicates the diesel

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fuel with NH3 has produced the maximum combustion temperature and heat transfer. The

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maximum cooling water availability of 8.9% fuel input is observed in B2 fuel corresponding to

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100% engine load with CR17.5 with NH5 ( Fig.3). This is because of higher chemical energy

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input during the operations and the progressive combustion [36]. The minimum cooling water

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availability of 55.4kJ/min was observed for B3 fuel at CR17.5, NH5 with 100% engine load.

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This is due to lower heat transfer to the engine cylinder walls by the rapid combustion process

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and lower LHV of fuel [37].

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Fig.4-6 presents the exhaust gas availability and destroyed availability (kJ/min) with

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different engine load, CR, and NH. The higher bael oil blended fuel operations engine produced

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about 40oC higher exhaust gas temperature as compared to diesel fuel mode [32]. This is due to

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late and inadequate combustion time for vegetable oil blended fuel [2]. The exhaust gas

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availability was lower for single hole nozzles compared to three and five hole nozzles. This is

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because of the lower temperature of combustion gas as results of lower combustion temperature

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[37]. Generally, higher compression ratio results in higher exhaust gas availability losses due to

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higher combustion temperature [38]. The maximum exhaust gas availability was observed in B3

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fuel in CR 17.5 with NH5 has 165.7kJ/min at 100% engine load (Fig.6). This is due to the long

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ignition delay of vegetable fuel and a higher combustion temperature in after burning

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combustion zone [3]. Because of these huge availability losses via exhaust gas, the efficiencies

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of blend fuel operations are lower than that of diesel mode [39]. The quantity of destroyed

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availability (Ades) is lower for the lower engine loads [40]. The minimum loads of 20% to 40%,

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lower combustion temperature causes lower destroyed availability [40]. The availability

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destruction is massive (as a percentage of fuel input) at no load conditions due to the zero shaft

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availability from the engine [22]. For NH1, the destroyed availability for vegetable oil blends is

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lower compared to neat diesel for lower CR. These values indicate that the destroyed availability

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was lower for vegetable oil blends or the input availability is utilized in other modes [19]. The

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minimum destroyed availability of 36.02% of fuel input is found for B1 fuel at the 100% engine

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load at CR16 with NH5 (Fig.6) which is 242.32kJ/min in neat diesel. For diesel fuel operation

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the minimum destroyed availability at 100% engine load is 37.36% of fuel input for CR17.5 with

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NH3 (Fig.5). It has been acknowledged that most of the exergy destruction is generated during

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the combustion process and a higher combustion pressure is beneficial for reduction of exergy

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destruction [20].

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Fig.7 shows the exergy efficiency for numerous test fuels with percentages of engine

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loads for different CR and NH. In lower engine loads, the exergy efficiency was lower for all test

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fuels. At no load condition, the useful shaft availability becomes zero and major input

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availability shared by the destroyed availability [19]. The effect of bael oil contents in blends

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affect the combustion process and cause for less cooling water and exhaust gas availabilities i.e.,

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higher the share of availability destroyed [39]. The maximum exergy value was found for B2

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fuel as 63.88% at NH3, CR17.5 with 100% engine load which is 61.79% for neat diesel. In the

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case of NH5, the maximum exergy efficiency values are slightly lower compared to NH3, this is

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due to the effect of charge dilution of fuels [34]. For B1 fuel the maximum exergy of 63.54% has

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been observed at NH5, CR16 with 100% engine load. For 75% engine load the maximum exergy

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efficiency has been observed as 61.01 % for B2 fuel at CR16 with NH3, which is 58.81% for

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neat diesel. For NH1, the maximum exergy efficiency was observed for neat diesel as 63.61% of

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fuel input availability for CR17.5 with 100% engine load. This demonstrates that blend fuel

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engine operations cannot be overlooked on the premise of their inferior thermal efficiency in a

322

diesel engine which turned into actually designed for the standard fossil diesel [19]. The

323

minimum exergy efficiency has been observed in neat diesel fuel as 26% of fuel input

324

availability corresponding to CR14, NH1 with no load condition. Thus compression ratio plays a

325

significant role in second law analysis and affecting combustion irreversibility through its effect

326

of gas temperature and pressure [22].

327

5.2 Brake specific energy consumption (BSEC)

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Fig.8 shows the influence of test fuels on BSEC for different values of compression ratio

329

and a number of nozzle holes. The BSEC is higher for numerous bael oil blended fuels when the

330

engine running at 25% engine loads and also lower compression ratio. This is due to inefficient

331

burning of fuel at lower CR [41]. At lower engine loads, the combustion process was

332

deteriorated due to the lower in-cylinder temperatures [42]. For NH1, increases the size of fuel

333

droplets into the engine cylinder because of large nozzle hole diameter and cause for inefficient

334

burning [43]. Thus the way the more amount of fuel injected into the combustion chamber to

335

produce the required power output [26]. BSEC of B1 fuels was almost same with neat diesel at

336

100% engine load at CR16, CR17.5 with NH1. The increasing of engine load will increase the

337

cylinder pressure and temperature which causes decreasing the tendency in the BSEC [44]. At

338

100% engine load, the B1 fuel has 1.2% lower BSEC than neat diesel in CR16 with NH3. At

339

100% engine load, the BSEC of B1 has been 3.3% lower with diesel in CR17.5 with NH5. The

340

vegetable oil blend fuel has higher BSEC compared to neat diesel for major operating conditions

341

[44]. Heating values are also lower for vegetable oil blends because of the lower calorific value

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of vegetable oil and DEE [17]. The longer ignition delay and lower burning velocity of vegetable

343

oil blends increase the BSEC due to the deterioration of the combustion process [10,45]. The

344

addition of DEE improves properties of vegetable oil such as reduces viscosity, auto ignition

345

temperature and improves cetane number and fire point [46]. For NH5 the BSEC is probably

346

lower compared to NH1 and NH3 due to the better atomization and air-fuel mixing [47]. The

347

minimum BSEC has been observed for B1 fuel as 12.45 MJ/ kW-hr at CR16 with 100% engine

348

load. For NH5, the B2 fuel has 1.3% lower BSEC compared to neat diesel at CR17.5 with 50%

349

engine load. From observation, the B2 and B3 fuel has more BSEC values needed for to produce

350

the equivalent diesel power output [44]. However, the BSEC values increase slightly at the

351

higher compression ratio due to dilution of fuel [34].

352

5.3 Brake thermal efficiency (BTE)

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The BTE of different fuel blends have been plotted towards percentage of engine loads as

354

shown in Fig.9. For any specified value of CR, the BTE will increases with increase in engine

355

load. The BTE is higher at higher loads because of the raised combustion temperatures inside the

356

cylinder and brake mean effective pressure [7]. Generally, NH1 is lower BTE values for the test

357

fuels compared to NH3 and NH5. The single injector hole could not produce the better

358

combustible mixture for combustion [43]. The BTE of diesel is higher compared to the bael oil

359

blends for most of the engine loads because of their high viscosity and higher fire point of bael

360

oil in lower CR [29]. The addition of DEE in the blends helps to reduce the bael oil viscosity and

361

auto-ignition temperature [15]. For NH3, the maximum BTE of 26.8% is observed in B2 fuel at

362

CR17.5 with 100% engine load which is 26.6% for neat diesel. The successful burning of

363

hydrocarbon species in the fuel creates to accomplish the peak thermal energy and decline the

364

fuel consumption quantity [16]. The effect of increasing number of nozzle holes increases the

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BTE for bael oil blends due to an account of the enhanced atomization leads to easy oxidation of

366

core of fuel droplets [47]. For B3 fuel, the BTE is lower compared to other test fuels at same

367

compression ratio and nozzle holes. This is due to more concentration of vegetable oils in the

368

blends leads to incomplete combustion [13]. From Fig.9, 100% engine load, the BTE of the neat

369

diesel fuel has been found as 29.2% at CR17.5 with NH5. The superior BTE has been observed

370

for B2 fuel as 29.32% at 100% engine load corresponding to CR17.5 with NH5; this is basically

371

because of the enhanced atomization, fuel spray character, and mixing of air-fuel, which brings

372

about better burning [47]. The lower self-ignition temperature of the DEE blends helps in

373

decrease surface tension produce complete combustion of the fuel and raising the BTE [12]. The

374

addition of DEE inside the bael oil and diesel, the peak cylinder pressure reduces. This is due to

375

the inferior heating value of the DEE blended fuel and the progressive combustion [34]. Low-

376

temperature combustion is relied upon to assistance the overall engine efficiency, principally as a

377

result of retarded cylinder thermal loss and probable of increasing combustion process from

378

weakening combustion which permits a greater amount of the energy to be extracted in the

379

working stroke [37]. At lower compression ratios the BTE was lower due to the lower peak

380

pressure, inefficient air/fuel mixing and incomplete burning of fuel [38].

381

5.4 HC emissions

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383

Fig.10. HC emission trends diminish for all fuels with an increment of engine loads due to

384

enhancement of combustion process [46]. A considerable reduction in HC emissions for B1 and

385

B2 fuel compared to neat diesel at 100% engine loads in CR17.5 with NH3. At 100% load, B1

386

and B2 blends have 15.6% and 8.5% lower HC emissions than the neat diesel at CR17.5 with

387

NH3. At 75% load, B1 blend has 10.5% lower HC emissions compared to neat diesel at CR16

17

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with NH3. This reduction happens due to the complete burning of the blended fuel with the

389

effect of DEE as to fulfill the combustion characteristics such as net HRR and mass-burn fraction

390

[49]. The HC molecules burned easily because of the elevated combustion temperature and

391

superior cylinder pressure at the working stroke, thus the way lower HC emissions results for

392

higher engine loads [26]. The maximum HC was observed in B3 fuel at 25% engine load was

393

23.8ppm at CR14 with NH3. Higher concentration of bael oil blends have a higher viscosity that

394

leads to incomplete combustion and increases hydrocarbon emissions [27]. For NH5, the

395

maximum HC emission observed as 20.3ppm for B3 fuel at CR14 and 100% engine load. This is

396

due to long ignition delay permits more time for the fuel vapors to diffuse and an excessive

397

fraction of fuel would be controlled inside the lean flame blow out region which causes an

398

increase in HC emissions [40]. The lower HC emissions for neat diesel at lower loads compared

399

to blended fuels [18]. However, the HC emissions slightly increase for NH5 with higher

400

compression ratio operations [34]. The impact of a number of nozzle holes will increase the HC

401

emissions due to excessive fuel-air ratio, lower combustion temperature, and shorter controlled

402

combustion period [47]. The vaporized fuel from the nozzle holes and sac volume enters the

403

engine cylinder at a smaller speed later in the cycle during expansion stroke and has little time to

404

mix with air when the combustion gas temperatures are nevertheless high [50,17]. This part of

405

the fuel remains generally unburned and is emitted in the exhaust [36]. Generally, the HC

406

emissions are a favor for following conditions NH3 with CR16, CR17.5, and NH5 with CR16.

407

5.5 CO emissions

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Fig.11 indicates that an impact of test fuels on CO emissions for different values of CR

409

and NH. Air-fuel proportion, injection pressure, engine load, type of fuel used and intake air

410

temperature influence CO emissions [4]. For NH1, the CO emissions are higher compared to

18

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NH3 and NH5. The primary factors are incomplete combustion of a fuel and inadequate air/fuel

412

mixing [6,43]. The maximum CO emission was observed in B2 fuel as 0.06% at 25% engine

413

load. The CO emission is lower for B1 fuel in most of the engine loads compared with B2 and

414

B3. The slow burning and large ignition delay period with lower engine load weakening the

415

oxidation of bael oil blends with air [5]. For NH3, the minimum CO emission was observed in

416

B1 fuel at 100% engine load with CR14. The better atomization and proper combustion of fuel

417

by the addition of DEE and lower concentration of bael oil in the blend result in lower CO

418

emission [15]. At 100% engine load, the B1 blend has been 20% lower CO emissions than the

419

neat diesel in CR15 with NH5. At the point, improved engine load and in-cylinder temperatures

420

enhance the combustion process and lower CO emissions [17]. At 100% load, B1 blend has been

421

20% lower CO emissions than the neat diesel corresponding to CR16 with NH5. The motive for

422

lower CO emission is due to stronger combustion with the consequence of DEE [51]. Then

423

again, the increased CO emission of 0.05% happens in an NH5 injector with CR17.5 with peak

424

engine loads. An increase of engine load results that combustion temperature increases and

425

dissociation environment for CO are greater favorable [18]. With further increment in load, the

426

CO emissions begin expanding once more and sharply upward push as still more fuel is injected

427

to increase the power output and observed that more CO emissions with higher compression ratio

428

[47].

429

5.6 NOx emissions

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Fig.12 shows the influence of test fuels on NOx emissions for different values of CR and

431

NH. Decrease in NOx emission is observed when the engine working by way of DEE blended

432

fuels with lower bael oil concentration [51]. Depending upon the burned gas temperature, the

433

numerous reactions of the kinetics of NO formation and henceforth the deviations is probably

19

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visible from the simple theoretical adiabatic flame temperature [52]. The development of NOx

435

begins by the side of the ignition gasses and the combustion close to chemically correct and lean

436

flammable products through premixed combustion phase [53]. For NH1, 100% engine load the

437

lower NOx was observed in B1 fuel at CR14. This is due to lower combustion temperature and

438

proper oxidization of hydrocarbon species in the fuel [15]. The maximum NOx emission has

439

been observed in B3 fuel as 444.5ppm at CR17.5 with 100% engine load. At higher bael oil

440

combination, NOx emission is elevated because of the increasing combustion chamber

441

temperatures at peak loadings [13]. At 100% engine load, NOx emission was 18.5% lower for

442

B1 fuel compared to neat diesel in CR17.5 with NH3 and B2 fuel has slightly lower than the neat

443

diesel at same operating condition. Similarly, B2 fuel blend has slightly lower NOx emissions

444

compare to neat diesel in CR17.5 with NH5 for 75% and 100% engine loads. On the off chance

445

that the excessive burned gas temperatures due to the presence of high turbulence mix rapidly

446

with colder air or air-fuel mixture, the reactions that deteriorate NO2 back to NO and oxygen are

447

frozen and reasonably higher concentrations of the NO2 result [56]. For NH3, the B1 fuel has

448

lower NOx emissions compared to other test fuels for the same compression ratio. The reduction

449

in peak combustion temperature was observed for test blends because of adding of the DEE,

450

which facilitates in diminishing the NOx emission [17]. From Fig.12, it may be visible that NOx

451

emission will increase with increasing the compression ratio for NH3. The mass of fuel

452

accumulated earlier than combustion process relies upon the ignition delay; lowers the ignition

453

delay lead to declining the peak temperatures and subsequently lowering the formation of NOx

454

[38]. The combustion duration is shortened for the blends due to the addition of DEE that helps

455

in NOx emission level [51]. Since the volume of injected particles of vegetable oil blend is

456

greater than the diesel fuel and the burning efficiency obviously reduced [54].

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457

5.7 Carbon dioxide The complete combustion of fuel in the combustion chamber is directly indicated by the

459

CO2 emission. The hydrocarbon fuels used in the combustion process, the CO2 and water vapors

460

are considered to be end products [15]. The complete combustion of fuel in the combustion

461

chamber is directly indicated by the CO2 emission [43]. From Fig.13, the CO2 emission initially

462

decreases, reach the lowest and subsequently increase with the increase in CR for all the test

463

fuels. It can be noted that the CO2 emission was lower for diesel fuel operation at CR14, single

464

hole nozzle and lower engine load. For NH1 operation, the fuel disintegration was not enough to

465

provide a time for complete evaporation and subsequent combustion due to fuel particle size

466

[43]. In the case of NH3, the CO2 emission has increasing trends for all test fuels. In NH3, B3

467

fuel has 20% higher CO2 emission compares to neat diesel at CR17.5 with 100% load. More CO2

468

is not much harmful to humans but it causes for ozone layer deflection and global warming [15].

469

It also depends on the exhaust gas temperature. The CO2 emission from biofuels combustion can

470

be absorbed by the plants and kept a constant level in the atmosphere [49]. The maximum CO2

471

emissions of 3.12% have been found in B2 fuel with five nozzle injector at 100% engine load

472

which is 2.81% for diesel fuel. But NH5 with CR 17.5 the CO2 emission level is lower compared

473

to the CR16. This may be due to higher combustion temperature about 1500oC the dissociation

474

of CO2 and H2O happens in the combustion chamber. The dissociation process reduces the CO2

475

level by CO2 to CO reduction [55]. The engine operated with lower load condition the CO2

476

emission for diesel fuel is lowest compared with oxygenated fuel blends [56]. At 75% to 100%

477

engine loads significant improvements in oxidation process due to higher combustion

478

temperature [55]. Throughout the engine operations, the CO2 emissions level increases

479

considerably with a number of nozzle holes and compression ratio [56].

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5.8 Smoke opacity

481

Smoke opacity is measured in percentage as shown in Fig.14. Either the inefficient fuel burning

482

or excess fuel is burning results higher smoke opacity. From Fig.14, smoke is considerably lower

483

for higher CR for all the test fuels. The heat of the compressed air is high at higher CR and lead

484

complete combustion [56]. At rated loads, the smoke formation is diminished because of an

485

oxygenated DEE that leads to complete combustion [57]. For NH1, the maximum smoke was

486

found in B3 fuel at CR14 and 100% engine load. At medium load conditions for all fuel has

487

lower smoke opacity. The use of vegetable oil blends in engines increases smoke opacity due to

488

ignition delay [6]. This increased ignition delay period is the reason for greater fuel accumulation

489

and increases the temperature during the power stroke and reduces the soot oxidation [42]. The

490

rapid progress of diffusion combustion may also favor for lowering the opacity in the range

491

between 10% and 25% than conventional diesel [58]. For NH3 with 100% engine load B1 fuel

492

has 42.3% with CR15 which is 44.7% for neat diesel. The oxygenated additive of DEE enhances

493

the combustion process and reduces the smoke intensity [17]. This may be due to increased

494

volatility, reduced viscosity and better mixing of air and fuel [15]. For NH5, the minimum

495

smoke opacity of 10.01% was observed in B1 fuel at CR17.5 with no load conditions. This lower

496

smoke is attributed by DEE in blends and better sprays formation of fuel by the nozzle holes

497

[14]. At 100% engine loads, B1 fuels have 9.2% lower smoke emission than diesel at CR15. At

498

CR17.5 the test fuel has higher smoke opacity for peak engine loads. The Soot emission is

499

generally formed in the region with a higher equivalence ratio and combustion temperature lies

500

between 1200oC to 2250oC [33]. The increasing of vegetable oils in the blends directly impacts

501

the smoke formation. This is due to poor atomization, insufficient time for oxidation and higher

502

molecular weight of vegetable oil [44].

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503

5.9 Cylinder pressure curve Fig.15 indicates the cylinder pressure variation against crank angle (-20o to 50o

505

connection with TDC) for CR17.5 with NH3. The maximum pressure relies depends on the

506

burned fuel fraction in the period of the premixed burning phase is the basis for compression

507

ignition engines [59]. The maximum pressure reduces when the addition of DEE in the bael oil

508

and diesel. The shorter ignition delay and wider spray pattern have formed because of lower

509

viscosity of the DEE in the blended fuel [60]. Optimum cylinder pressure rise was observed due

510

to a reduction in the premixed fuel and the lower LHV of the DEE blends [60]. The peak

511

pressure rate principally relies on the rate of combustion in the initial stage and fuel taking part

512

in the uncontrolled heat release phase [35]. The rate of pressure has minimum and the pressure

513

raise increases with increasing applied load to the engine [29]. In case of blend B2, the

514

combustion starts nearly the same crank position, which may reflect the pressure rise like that the

515

neat diesel. For B3 fuels, the rate of pressure rise is lower compared to neat diesel and B1fuels

516

due to the lower LHV and inefficient burning of fuel [17].

517

5.10 Maximum cylinder gas temperature

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Average maximum cylinder temperature against crank angle (-20 to 70o reference to

519

TDC) was plotted in Fig.16. At 25% to 100% load, the maximum cylinder gas temperature is 3%

520

lower for B1 blends compared to neat diesel. In 100% engine load, the average maximum

521

cylinder temperature is 2% lower for B2 fuel compared with neat diesel at CR17.5 with NH3. In

522

75% engine load, the maximum cylinder temperature is 1% lower for B2 fuel compared with

523

neat diesel at CR17.5 with NH3. The reduced maximum combustion temperature was a reason

524

for reducing NOx emissions [52]. Low-temperature combustion is predicted to benefit overall

525

engine efficiency, primarily because of reduced cylinder heat loss and potential of molecular

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properties of the expanding combustion gasses from dilute combustion phase to allow a greater

527

amount of the energy to be extracted in the expansion stroke [60]. However, the thermal

528

efficiency little bit decreased according to the Carnot theorem that is the efficiency based on the

529

temperature limits [53].

530

5.11 Average Heat Release Rate

RI PT

526

The heat release rate was calculated by the first law of thermodynamics based on the data

532

of the recorded cylinder pressure [34]. From the equation, the heat released by combustion of

533

fuel is the sum of net heat release rate and heat absorbed by the cylinder wall. The heat release

534

rate calculated by the following equation (1): KL

KM

N

KR

P

KT

= NOP +Q KM/ + NOP +S KM/ + Where

536

KL

KM

CUV CW

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531

(1)

is the amount of heat transfer rate (J/deg CA), θ is crank angle (deg), X is ratio

of specific heats, P is cylinder pressure (N/m2), V is the instantaneous volume (m3) of the

538

cylinder and YZ is the cylinder wall heat losses [34]. This calculation was done by assuming

539

uniform properties throughout the combustion chamber and the entire mixture is an ideal gas

540

where the specific heats are dependent on temperature [61]. Where

542

CUV CW

CUV CW

is the amount of heat transfer rate to the engine cylinder wall eq. (2).

= ℎ] × ^ × _ − Z 

(2)

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543

Where ℎ] is the convective heat transfer coefficient between cylinder wall and combustion gas,

544

A is the convection heat transfer area, _ is the combustion gas temperature and Z is the

545

cylinder wall temperature.

546

Hohenberg’s correlation was used for calculating instantaneous heat transfer [34]. The

547

instantaneous heat transfer coefficient across the walls for the engine was estimated using eq. (3)

24

ACCEPTED MANUSCRIPT

548

P` ×abc.e ×fg hP.ic.e

ℎ] = 1

R c.cj ×klc.m

6

(3)

Fig.17 shows the variations of average heat release rate (HRR) as a function of crank

550

angle (CA) position (-20 and 40o CA) for CR 17.5 with NH3 for diesel and bael oil blends. We

551

observed the negative heat release for all fuels because of the combustible mixture express

552

cooling effect as a result of vaporization of the accumulated fuel during the ignition delay period

553

in compression stroke [16]. As soon as the combustion phenomenon starts the HRR gets to be

554

distinctly positive [60]. One can observe that for each load considered, the average heat release

555

curve for the B2 blend lies, at the beginning, a little lower than the corresponding one for diesel

556

fuel. After the ignition delay period, premixed air/fuel mixture burns quickly, followed by the

557

diffusion combustion phase. At this point, the HRR is controlled with the aid of the charge of

558

air/fuel mixing [51]. The rate of HRR has been higher for neat diesel operation due to the higher

559

LHV compared to blended fuel [16]. At lower engine loads, the rate of HRR has been lower due

560

to the smooth burning of a small quantity of fuel [34]. For higher engine loads more amount of

561

fuel injected into the cylinder to produce higher output power by rapid combustion process [60].

562

6. Uncertainty analysis

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549

The experimental analysis is affected by some uncertainties due to the instruments

564

imperfections and the inaccuracy of fitting constant determination. It is used to estimate the

565

experimental error [62].

Calculation of uncertainty in brake power

566

567

noT oT

nq r

nt r

nu r

= p q ' +  t ' +  u '

Calculation of uncertainty in fuel consumption or fuel power

568

569

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563

nvw vw

nR r

nk r

= p R ' +  k '

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Calculation of uncertainty in BTE

570

571

nokx okx

noT r

= p

oT

' +

nvw r vw

'

The instruments for measurements are chosen with a view to keeping the experimental

573

uncertainties as minimum as possible. The probable errors in the stopwatch (∆t), speed indicator

574

(∆N), graduated burette (∆V), measuring scale (∆R), and strain gauge load type load cell (∆W)

575

are 0.01 sec, 5rpm, 0.001m and 0.1 kg respectively. Uncertainties of some measured and

576

calculated parameters are given in Table 6. Based on the above values, the calculated engine

577

performance is believed to be accurate within ±3% [45].

578

7. Conclusion

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572

Taking interested in consideration of the need for alternate fuels, the experimental study

580

was completed in order to work the DI diesel engines by way of bael oil and diesel with DEE

581

additive. From the results of the comparative evaluation, the diesel and vegetable oil blend with

582

DEE ignition enhancer has the potential to reduce the harmful pollutions further more enhance

583

the engine performance and combustion. As an end result, fuel mixture B2 (60% diesel: 30%

584

bael oil: 10% DEE) offered the phenomenal performance while compared towards other blends

585

in expressions of brake thermal efficiency, carbon monoxide, hydrocarbon emission and oxides

586

of nitrogen.

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587

•

588

CR17.5 with NH5 is 677.8kJ/min. The cooling availability in the range of 8-12% of fuel input is

589

accessible in blended fuel operations compared to about 7-10% to that of fossil diesel. The

590

maximum cooling water availability of 8.9% fuel input is observed in B2 fuel at 100% engine

591

load corresponding to CR17.5 with NH5.

The maximum input availability has been found in B1 fuel at 100% load corresponding to

26

ACCEPTED MANUSCRIPT

592

•

593

NH5 has 165.7kJ/min at 100% engine load. The maximum availability destroyed in D fuel has

594

been 73.43% of fuel input in CR14 with NH1 at zero engine loads. The minimum destroyed

595

availability of 36 % of fuel input was found for B1 fuel at the 100% engine load in CR16 with

596

NH5. Diesel fuel mode shows that the minimum destroyed availability of 37.36% of fuel input

597

at 100% engine load in CR17.5 with NH3.

598

•

599

engine load in CR17.5 with NH3, which is 62.79% of fuel input for neat diesel. Be that as it

600

may, on account of utilizing diesel fuel the minimum exergy efficiency is observed as 26.72%

601

of fuel input at no load corresponding to CR14 with NH1.

602

•

603

with 100% engine load. The higher BTE of 29.32% for B2 blends when the engine running

604

100% engines load corresponding to CR17.5 with NH5.

605

•

606

compared to neat diesel at CR17.5 with NH3. The maximum HC is observed in B3 fuel and

607

25% engine load was 23.8ppm at CR14 with NH3. At 100% load, B1 blend has been 20% lower

608

CO emissions compared to neat diesel corresponding to CR16 with NH5.

609

•

610

injector at 100% engine load which is 2.81% for diesel fuel. At 75% to 100% engine loads

611

significant improvements in oxidation process and produce more CO2 species. At 100% engine

612

loads, B1 fuels have 9.2% lower smoke opacity than diesel at CR15 with NH5. The B2 fuel has

613

slightly lower smoke opacity compared to neat diesel for 100% engine loads at CR16, CR17.5

614

with NH5.

RI PT

The maximum exhaust gas availability has been observed in B3 fuel at CR 17.5 with

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The maximum exergy value has been observed 63.88 % of fuel input for B2 fuel at 100%

TE D

The minimum BSEC has been observed for B1 fuel as 12.45 MJ/ kW-hr at CR16, NH5

EP

At 100% engine load, B1 and B2 fuels have 15.6% and 11.5% lower HC emissions

AC C

The maximum CO2 emissions of 3.12% have been found in B2 fuel with five nozzle

27

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615

•

616

diesel in CR17.5 with NH3 and B2 fuel has slightly lower than the neat diesel at the same

617

operating condition. The B2 fuel blend has slightly lower NOx emissions compared to neat

618

diesel in CR17.5 with NH5 for 75% and 100% engine load conditions. A trade off zone of

619

engine loads (65% and 75% of the maximum load) was recognized regarding the gas emissions,

620

technical constraints of engines and exergy efficiency for the most favorable performance of the

621

CI engine.

622

References

623

[1] Bruno Alessandro D, Gianni Bidini, Mauro Zampilli, Paolo Laranci, Pietro Bartocci. Straight

624

and waste vegetable oil in engines: review and experimental measurement of emissions, fuel

625

consumption and injector fouling on a turbocharged commercial engine. Fuel 2016; 182:198-

626

209. http://dx.doi.org/10.1016/j.fuel.2016.05.075

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At 100% engine load, the NOx emission is 18.5% lower for B1 fuel compared to neat

[2] Peter Emberger, Dietrich Hebecker, Peter Pickle, Edgar Remmele, Klaus Thuneke. Emission

628

behaviour of vegetable oil fuel compatible tractors fuelled with different pure vegetable oils.

629

Fuel; (accessed Dec 2016). http://dx.doi.org/10.1016/j.fuel.2015.11.0751

TE D

627

[3] Sarveshwar Reddy M, Nikhil Sharma, Avinash Kumar Agarwal. Effect of straight vegetable

631

oil blends and biodiesel blends on wear of mechanical fuel injection equipment of a constant

632

speed

633

http://dx.doi.org/10.1016/j.renene.2016.07.072

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diesel

engine.

Renewable

Energy

2016;99:1008-1018.

634

[4] Ramkumar S, Kirubakaran V. Biodiesel from vegetable oil as alternate fuel for CI engine and

635

feasibility study of thermal cracking: A critical review. Energy Conversion and Management

636

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Thermal

Engineering

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EP

810

jatropha

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812

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811

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821

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TE D

828

M AN U

824

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830

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831

Nomenclature ASTM B1 B2

AC C

832

EP

829

American society for testing and materials 70% diesel+ 20% bael oil+ 10% DEE 60% diesel+ 30% bael oil+ 10% DEE

B3

50% diesel+ 40% bael oil+ 10% DEE

BSEC

brake specific energy consumption

BTE

brake thermal efficiency

37

ACCEPTED MANUSCRIPT

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

L

exergy loss

LHV

lower heating value

NH

number of nozzle holes

NOx

oxides of nitrogen

SVO

straight vegetable oil

VCR W

SC M AN U

TE D

EP

AC C

TDC

RI PT

CA

top dead centre

variable compression ratio Watt

Greek letter γ

ratio of specific heats

ɳII or ɛ exergy effieiency 38

ACCEPTED MANUSCRIPT

θ

crank angle

Ρ

density

ʋ

viscosity (centi stroke)

dead state condition

a

atmosphere condition

cw

cooling water

des

Destroyed

eg

exhaust gas

in

input

M AN U

SC

0

RI PT

Subscripts

AC C

EP

TE D

833

39

ACCEPTED Table 1 Properties of diesel, bael oil and DEEMANUSCRIPT Diesel C16H34 830 2.7 200-400 50 -20 180-330 42800 14.9

Bael oil C18H36O2 896 24.3 <370 51.7 -5 298 36300 12.4

Table 2 Properties of blended fuels

B1 B2 B3

Density (kg/m3) at 32 0c

SC

1 2 3

Kinematic viscosity (cS) 6.81 7.99 10.11

Blend

831 849 876

M AN U

S.No

DEE C2H5OC2H5 713 0.23 160 >124 -110 35 33900 11.1

RI PT

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

Calorific value (kJ/kg) 41218 40476 39734

Table 3 Technical specifications of the test engine Type

AC C

EP

TE D

No. of cylinders/ No. of strokes Rated power Bore (mm)/Stroke(mm) Type of ignition Compression ratio Injection pressure (standard) Injection timing (standard) Speed Diameter/ no. of nozzle hole Dynamometer

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 22o bTDC 1500 Rev/min 0.8mm/ 1; 0.3mm / 3; 0.2mm/5 Eddy current dynamometer; Water cooled; ModelTMEC10; RPM 1500-6000; Make - Technomech Pvt., Ltd. Piezo electric sensor; Model –M111A22; Resolution0.1psi; Sensitivity – 1mV/psi. Make- Kistler; Model- 2614C11; Speed range - 0 to 12000 rpm; Crank angle signal - 720* 0.5 degree. USB-6210; 16AI; 4DI; 4DO USB- multifunction I/O devive; Make- National instruments. Differential pressure transmitter; Make- Broiltech; Model- FCM. Make- Sensortronics; Model-60001. Make-Wika; Model- SL1.

Cylinder pressure sensor Crank angle encoder Data acquisition system Fuel flow measurement Load cell Air flow measurement

1

ACCEPTED MANUSCRIPT Table 4 Technical specifications of gas analyzer AVL DI444

Oxides of nitrogen (NOx) Carbon dioxide (CO2)

Measuring range: 0-10% vol

Resolution: 0.01% vol

0-20000 ppm vol 0-5000 ppm vol

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

0-20% vol

0.1 % vol

Absorption

Measuring Range Accuracy & repeatability

0-100%

0-99.99m-1

±1% of full Scale

Better than ±0.1 m-1

Resolution

0.1%

0.01 m-1

Rpm

Oil Temperature

400-6000 1/min ±10

0-150º C

±1

±1ºC

M AN U

Opacity

SC

Table 5 Technical specifications of smoke meter AVL 437C

Accuracy: < 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.

RI PT

Measured quantity: Carbon monoxide (CO) Hydrocarbon (HC)

Table 6 Uncertainties of some measured and calculated parameters

TE D

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

EP

Parameter NOx CO HC CO2 Smoke opacity Kinetic viscosity BTE BSEC

AC C

S.No 1 2 3 4 5 6 7 8

2

±2ºC

M AN U

SC

RI PT

ACCEPTED MANUSCRIPT

AC C

EP

TE D

Fig.1. Input availability, cooling water availability for NH1

Fig.2. Input availability, cooling water availability for NH3

M AN U

SC

RI PT

ACCEPTED MANUSCRIPT

AC C

EP

TE D

Fig.3. Input availability, cooling water availability for NH5

Fig.4. Exhaust gas availability, destroyed availability for NH1

M AN U

SC

RI PT

ACCEPTED MANUSCRIPT

AC C

EP

TE D

Fig.5. Exhaust gas availability, destroyed availability for NH3

Fig.6. Exhaust gas availability, destroyed availability for NH5

M AN U

SC

RI PT

ACCEPTED MANUSCRIPT

TE D

Fig.7. Exergy effificency Vs load

AC C

EP

Fig.8. Brake specific fuel consumption Vs load

Fig.9. Brake thermal efficiency Vs load

M AN U

SC

Fig.10. Hydrocarbon Vs load

RI PT

ACCEPTED MANUSCRIPT

AC C

EP

TE D

Fig.11. Carbon monoxide Vs load

Fig.12. Oxidies of nitrogen Vs load

Fig.13. Carbon dioxide Vs load

TE D

M AN U

SC

Fig.14. Smoke opacity Vs load

RI PT

ACCEPTED MANUSCRIPT

AC C

EP

Fig.15. Cylinder pressure Vs crank angle

Fig.16. Cylinder gas temperature Vs crank angle

SC

RI PT

ACCEPTED MANUSCRIPT

AC C

EP

TE D

M AN U

Fig.17. Heat release rate Vs crank angle

ACCEPTED MANUSCRIPT

Highlights; Feasibility of using aegle marmelos blend in Compression Ignition (CI) engine.

•

Combined effect of cylinder compression ratio and number of nozzle holes.

•

Performed the exergy analysis to diagnoses the various losses of availabilities.

•

The maximum exergy efficiency of B2 fuel has been found 63.88 % of fuel input.

•

CO and BSEC improved with the 5 holes nozzle.

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