An experimental study of combustion, performance, exergy and emission characteristics of a CI engine fueled by Diesel-ethanol-biodiesel blends

Accepted Manuscript An Experimental study of Combustion, Performance, Exergy and Emission characteristics of a CI engine fueled by Diesel-Ethanol-Biodiesel Blends

Abhishek Paul, Rajsekhar Panua, Durbadal Debroy PII:

S0360-5442(17)31659-6

DOI:

10.1016/j.energy.2017.09.137

Reference:

EGY 11631

To appear in:

Energy

Received Date:

21 July 2017

Revised Date:

25 September 2017

Accepted Date:

28 September 2017

Please cite this article as: Abhishek Paul, Rajsekhar Panua, Durbadal Debroy, An Experimental study of Combustion, Performance, Exergy and Emission characteristics of a CI engine fueled by Diesel-Ethanol-Biodiesel Blends, Energy (2017), doi: 10.1016/j.energy.2017.09.137

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ACCEPTED MANUSCRIPT

Highlights

  

Exergy analysis of an existing diesel engine fuelled with various Diesel-EthanolBiodiesel blends. Optimal performance, emission & exergy characteristics achieved with Diesel-EthanolBiodiesel blend with 15% ethanol. Possible frame work for a sustainable use of ethanol in a CI engine.

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An Experimental study of Combustion, Performance, Exergy and Emission characteristics of a CI engine fueled by Diesel-Ethanol-Biodiesel Blends.

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Abhishek Paula*, Rajsekhar Panuaa, Durbadal Debroya.

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Mechanical Engineering Department, NIT Agartala, Barjala,Jirania-799001.

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*corresponding [email protected]

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Abstract

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The present work is an attempt to conduct a complete analysis of a CI engine subjected to a number of blends of Diesel-ethanol and Pongamia piñata methyl ester (PPME). In this study, the PPME percentage is fixed at 50% and ethanol percentage is increased from 5% to 20% with intervals of 5%, thus reducing the diesel participation. A comprehensive analysis of performance, exergy, combustion and emission characteristics was carried out, which lead to a conclusion that the D35E15B50 blend with 15% ethanol showed best engine performance characteristics with 21.17% increase in brake thermal efficiency and 4.61% decrease in BSEC at full load. The combustion analysis also revealed increase in cylinder pressure and heat release rate indicating improvement in combustion condition for the above-mentioned blend. The D35E15B50 blend also showed a substantial improvement in unburned hydrocarbon and carbon monoxide emissions but it was penalized with a marginal increase in NOx emission. The exergy analysis showed a 25.64% increase in exergetic efficiency and 22.02% decrease in exergy destruction rate and 21.06% decrease in entropy generation rate at full load condition for D35E15B50 blend. The tradeoff study involving BSEC, NOx emission and sustainability index indicated a higher sustainability prospect for the D35E15B50 .

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Keyword- Exergy destruction rate, Sustainability index, Entropy Diesel-Ethanol-Biodiesel, Diesel Engine.

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Nomenclature

1

8 9 10 11 12 13 14 15 16 17 18 19 20

PPME

Pongamia piñata methyl ester

T5,out

BSEC

Brake Specific Energy Consumption.

Tex,I

NOx

Oxides of Nitrogen.

Tex,O

Carbon Monoxide Carbon Dioxide high-oleic methyl ester Exergy transfer rate of intake air. Exergy transfer rate of fuel. Exergy transfer rate of exhaust gas. Exergy transfer rate of shaft work. Exergy transfer rate due to heat transfer rate to coolant

T6 𝑚1 𝑚2 𝑚5 𝑚𝑤,3 CP1 Φ qLHV

CO CO2 HOME 𝐸1 𝐸2 𝐸3 𝐸4 𝐸5

Outlet temperature of cooling water from the engine. Exhaust gas temperature at inlet of the calorimeter. Exhaust gas temperature at outlet of the calorimeter. Average engine body temperature. Mass flow rate of air. Mass flow rate of fuel. Mass flow rate of cooling water. Mass flow rate of water in the calorimeter. Constant pressure specific heat of air. Chemical exergy factor. Lower calorific value of the fuel

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𝐸6 𝐸𝑑𝑒𝑠 𝐸𝑝ℎ 3 𝐸𝑐ℎ 3 T0 P0 T1 P1 T3 P3 TC,I TC,O T5,in

Exergy transfer rate due to heat transfer rate to ambient. Exergy destruction rate. Physical exergy of exhaust gas. Chemical exergy of exhaust gas. Dead state temperature. Dead state pressure. Intake air temperature. Intake air pressure. Exhaust gas temperature. Exhaust gas pressure. Water temperature at inlet of calorimeter. Water temperature at outlet of calorimeter Intel temperature of cooling water in the engine.

R CP3 CP,w 𝑛 xi єi ω τ 𝑄6 𝜑 SI 𝑆𝑔𝑒𝑛

Gas constant. Constant pressure specific heat of exhaust gas. Constant pressure specific heat of cooling water. mole number per kilogram of exhaust gas the molar fraction of each of the component standard chemical exergy of each component Angular velocity of engine. Engine torque. Heat loss to the environment. Exergetic Efficiency. Sustainability index. Entropy generation rate.

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Introduction

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The global concern regarding the energy insecurity and consistent obligations of complying with ever-stringent emission norms have diverged the attention of fuel industry from fossil based fuels to new and renewable fuel sources for IC engine application. The necessity to reduce the carbon foot print by means of eco-friendly and non-conventional fuels have compelled the researchers to explore different types of fuels and additives along with new and revolutionary fuel injection techniques. Diesel engines have become the prime choice not only for the heavy transportation sector but also for power generation sector as cogeneration system owing to its higher efficiency and greater power output [1-3]. However, constantly converging emission norms have obliged the use of Diesel engines, as it is very much culpable for high emissions of NOX, particulate matters and other polycyclic aromatic hydrocarbons. One way to overcome this situation is incorporate oxygen releasing molecules in fuel. There has been a substantial number of research works on rationalizing the usage of biofuels such as biodiesel, Alcohols etc., rich in reactive oxygen molecules in their structure, as a sustainable, renewable and viable alternative for existing diesel power trains without the necessity to adopt radical changes in the existing diesel platforms.

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There are a growing number of works devoted towards production and use of different biofuels. At present, the most common alternative to Petroleum Diesel is biodiesel, chemically known as alkyl ester of fatty acids from different plant and animal sources. Biodiesel is a renewable and sustainable fuel that can be produced from any kind of edible and non-edible vegetable oils, waste vegetable oils, animal fats etc. Due to its organic nature, biodiesel has 10-11% oxygen in its structure that allows widespread oxidation of fuel, resulting in more complete combustion [4, 5]. Due to its higher cetane number it has good self-ignition characteristics, which aids in it combustion and allows achieving similar values of thermal efficiency in comparison with engine powered by diesel fuel [6]. Biodiesel also reduces the formation of CO, HC, and soot particles in engine [7-10]. Biodiesel also decreases the lifecycle emission of CO2 [11]. However, the NOx emission from the engine with

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biodiesel is found to be higher [12,36]. This increased NOX emission from biodiesel fueled engines can be countered by incorporating ethanol into the fuel blend. Additionally, due to their highoxygenated content, ethanol can also have the additional effect of reducing the emission of hydrocarbon, CO and smoke of a CI engine. Ethanol has an oxygen content of 34% and a latent heat of vaporization of 840 kJ/kg [13]. These encouraging aspects of ethanol have been well documented by a number of researchers over the year. Irshad Ahmed [14] studied the effect of ethanol Diesel blends on the performance and emission of a CI engine. The results of this study reported a 41% reduction in particulate emission, 27% reduction in CO, and 5% reduction in NOx emissions. Li et al. [15] studied the effect of ethanol on CI engine combustion and found that simultaneous reduction in smoke, NOx and CO emissions is possible with 15% ethanol blended with Diesel. Hardenberg and Schaefer [16] studied the effect of ethanol on a CI engine. The NOx and smoke emission reducing effect of ethanol was again reported in this study. Based on such encouraging results some previous works conducted by the same researchers [3,17,18] also concluded with reduced emissions of NOx, CO and smoke with marginal increase in performance of the engine using ethanol as an additive. The same researchers also found that a high percentage of biodiesel, as high as 50% by vol. can be to stabilize a Diesel-ethanol-biodiesel blend with high ethanol content in it [18]. The exergy and energy analysis of IC engine have been discussed for almost two decades as a tool to assess the different losses occurring during the operation of an IC engine. Exergy represents the amount of useful work that a system can provide when it moves toward the reference environment by a reversible process [19,20]. Exergy analysis can be used to determine the type, location and magnitude of energy losses in different parts of an engine. Hence, exergy analysis opens a window of opportunity to take necessary measures for reducing the losses in different parts of engine. Various investigators conducted a number of studies on exergy analysis of IC engine. Azoumah et al [21] optimized the performance of a DI engine using biofuels by means of exergy analysis. The results showed that the amalgamation of exergy analysis and gas emission analysis is a very effective tool for optimal application of loads on a CI engines. Benjumea et al. [22] applied exergy analysis to a turbocharged (TC) automotive diesel engine fuelled with neat palm oil biodiesel (B100) and diesel fuel (B0). Tests were conducted under steady state operating conditions, at two altitudes above sea level: 500 and 2400 m. It was found from the study that exergy destruction is increased sharply when combustion is started, indicating that this process was the main source of irreversibility. Caliskan et al [23] investigated the effect of varying dead state temperatures on exergy efficiency of a high-oleic methyl ester (HOME) fueled internal combustion engine. It was found that the exergetic efficiency increased with the decrease of dead state temperature. López et al. [24] compared the exergetic performance of a DI diesel engine using olive–pomace oil methyl ester, mineral diesel, and their mixtures at full load and at different speeds. The exergetic performance parameters of the applied fuels were found to be identical to base diesel operation. Based on the exergy results, it was resolved that olive – pomace oil biodiesel and its blends with diesel fuel can substitute the use of diesel fuel in CI engines without any exergy cost increment.

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The present work is an attempt to study the effect of increasing ethanol percentage in Dieselbiodiesel blends by means of in-depth study of combustion, performance, emission and exergy parameters of an engine. In this study, the biodiesel percentage was fixed to 50% and ethanol percentage was increased from 5 to 20% in 5% incrementing intervals, thus reducing the Diesel percentage subsequently. The results of such analysis could be of great interest to researchers in determining optimum fuel blends and engine operating parameters to achieve more economical and eco-friendly operations.

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2 Experimental Setup and Procedure

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2.1 Test Fuels and their Properties.

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The base fuels used in this experimental work are High speed Diesel, Pongamia piñata methyl ester (PPME) and Ethanol. The Diesel and Ethanol are collected from local fueling stations and local chemist shop respectively. The PPME (Commercially known as Biodiesel) is produced by transesterification of Pongamia pinata seed oil and the whole process is carried in the IC engine laboratory of NIT Agartala. Previous research done by the same researchers showed that ethanol has limited miscibility upto 10% (v/v) in Diesel [25]. It was also witnessed that miscibility of ethanol can be increased by introducing PPME into the blend and a minimum of 45% and 50% PPME is required to stabilize a blend containing 15% and 20% ethanol [18]. Hence, in this experimental work the biodiesel percentage is fixed at 50% by volume. Then the ethanol in introduced into the blend in 5% by volume concentration. Subsequently the Diesel percentage was reduced to 45%. In subsequent blends, the ethanol percentage is increased by 5%, thus producing blends with 10%, 15% and 20% ethanol. These blends are named as D45E5B50 (Containing 45% diesel, 5% ethanol and 50% PPME), D40E10B50 (Containing 40% diesel, 10% ethanol and 50% PPME), D35E15B50 (Containing 35% diesel, 15% ethanol and 50% PPME) and D30E20B50 (Containing 30% diesel, 20% ethanol and 50% PPME). These blends are shown in Fig-1. The properties of the base fuels are shown in Table-1 [3].

89 90 91 92 93 94

99 100 101 102 103 104 105 106 107 108 109 110 111 112 113

Property Diesel Ethanol Biodiesel Density (kg/m3 ) 820 789 886 Kinematic Viscosity(cSt) 2.51 1.09 4.06 Calorific Value (kJ/kg) 42650 26950 35866 Flash Point ( °C) 52 12.77 217 Fire Point (°C) 64 13.5 220 Cetane Number 48 7 55 114 115

Fig-1: Tested diesel-biodiesel-ethanol Blends

Table:1: Properties of the Base Fuel

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2.2 Experimental Setup and methodology

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The experimental work has been conducted on a single cylinder, 4 stroke, water cooled, naturally aspirated, stationary DI engine that can produce a maximum brake power of 3.6 kW at 1500 RPM. The engine has a compression ratio of 18:1. An eddy current dynamometer (Model-AG10, Make Saj test plant Pvt. Ltd.) is used to apply load to the engine. The load cell used along with the dynamometer is a 'S' type universal load cell (Model-60001, Make: -Sensortronics ). The dynamometer can cope up to a speed of 10000 RPM. A Kublar DE1 8.3700.1321.0360 Crank angle sensor has been fitted to the crankshaft of the engine to measure the engine rotational speed by synchronizing it with the rotation of the crankshaft. The crank angle sensor can measure engine rotational speed with 1 degree precision. A piezoelectric pressure sensor (Model-HSM111A22, make-PCB Piezotronics, INC.) is installed in the cylinder head to measure the cylinder pressure. The piezoelectric pressure transducer is a polystable quartz crystal, which produces an electric charge that is proportional to the pressure developed in the cylinder. A piezo powering unit (ModelAX-409, make- Cuadra) powers the transducers. A differential pressure transducer (ModelEJA110A-DMS5A-92NN, make- Yokogawa) used to measure the liquid fuel flow into the engine. The flow rate of fluid is calculated by measuring the difference of fluid pressure in the pipe. Differential pressure flow meters have a primary and a secondary element. The primary element is designed to produce a difference in pressure as the flow increases. The present experimental set up uses a burette tube of 12.4mm diameter for this purpose. The indicted air is routed through an air surge tank fitted with an orifice meter, a manometer and a pressure transducer (Model-SL-1, make WIKA) to measure the flow rates and the associated drop values needed for the computation of the volumetric efficiency of the engine under different operating modes. Two types of temperature sensors are used for measuring temperature at different point of the engine test bed. The high temperatures of the flue gas at the exhaust manifold before and after the calorimeter is measured by two K-type thermocouples as these thermocouples can handle temperatures up to 1260°C. Three more resistance temperature detectors (RTDs) are connected at the inlet water at pump outlet, the Engine outlet water on the engine head and calorimeter water outlet. All the instruments are interfaced to a computer through a centralized data acquisition system synchronized to the crank angle encoder. GUI-based post-processing software “Engine soft” (Developed by Apex Innovation Pvt. Ltd.) was used to analyze the data stream.

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The emission from the engine primarily consists of carbon dioxide, carbon monoxide, oxides of nitrogen, unburned hydrocarbon , water vapour etc. An AVL Digas 444 5-gas analyzer is used to measure the emissions of NOx, unburned hydrocarbon (UBHC), CO, CO2, and O2. The emissions of CO, CO2, and O2 are measured in terms of volume percentage. The emission of UBHC and NOx is measured in terms of ppm (Vol.). CO, CO2, and UBHC emissions are measured on the basis of NonDispersive-Infrared (NDIR) detection principle, while NOx and O2are measured by means of precalibrated electrochemical sensors. The major details of the engine is given in Table -2 and the circuit diagram of complete experimental set up is shown in Fig-2.

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The experiments are carried out at 20%, 40%, 60% , 80% ,100% and 120% load conditions. The engine is first operated on all the above-mentioned load conditions with Diesel to get a baseline data set. Before testing a new fuel sample, the engine worked for sufficient time to consume the remaining fuel from the previous experiment. Special care is taken to keep a constant speed of the engine (±10rpm) during data acquisition for each case of engine operation at different loads. The authenticity of the reading increased by taking six consecutive readings at the same condition and averaging them. The whole experimentation is done at an ambient temperature of 30-31°C and at a relative humidity of 70%.

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Fig-2: Complete Experimental Setup Engine type Bore and stroke Max. power Compression ratio Swept volume Fuel injection pressure Operating Speed Injection Timing

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Kirloskar, Model TV-1, 4 stroke Water cooled, VCR Engine. 87.5 mm and 110mm 3.6 kW(1500 RPM) 18 661 cc 205 BAR 1500 RPM. 23O BTDC

Table-2: Specifications of Test Engine

2.3 Theoretical considerations

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The complete thermodynamic analysis of the working of the engine was done based on a number of postulates and assumptions. It was assumed that the engine was operated at steady state condition and the intake air and the exhaust gases behaved like ideal gases. Further, the kinetic and potential exergies of the engine were neglected, as they didn’t pose any significant effect on the total exergy of the engine.

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The total exergy balance equation of the engine as a control volume can be written as,

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𝐸1 + 𝐸2 = 𝐸3 + 𝐸4 + 𝐸5 + 𝐸6 + 𝐸𝑑𝑒𝑠……………………. (1)

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Where, 𝑬𝟏 , 𝑬𝟐 , 𝑬𝟑 and 𝑬𝟒 are exergy transfer rate of intake air, intake fuel, exhaust gas and shaft

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work respectively. 𝑬𝟓 and 𝑬𝟔 are exergy transfer rate due to heat transfer rate to coolant and

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ambient respectively. 𝑬𝒅𝒆𝒔 is the exergy destruction rate.

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The exergy transfer rate of intake air is given by,

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[ [

( )] + 𝑅𝑇 𝑙𝑛( )]

𝐸1 = 𝑚1 𝐶𝑝1 𝑇1 ‒ 𝑇0 ‒ 𝑇0𝑙𝑛

𝑇1

𝑃0

0

𝑇0

𝑃1

……………… (2)

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Where, 𝒎𝟏 , Cp1, T1 and P1 are the mass flow rate, constant pressure specific heat, temperature and pressure of intake air respectively. R is the gas constant (0.287 kJ/kg K) and T0 and P0 are dead state temperature given as 31° C (304 K) and 101.315 kPa respectively.

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The specific heat Cp1 of the intake air can be found by the following equation [26],

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𝐶𝑝1 = 1.04841 ‒ 0.000383𝑇1 +

9.45378𝑇12 107

‒

4.49031𝑇12 1010

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1014

……………. (3)

190 191

+

7.92981𝑇12

The fuel exergy includes the chemical exergy only, which is measured by the following equation [27],

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𝐸2 = 𝑚2𝛷𝑞𝐿𝐻𝑉 ………………………………….. (4)

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Where, m2 and qLHV are the mass flow rate and lower calorific value of the fuel. Φ is a chemical

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exergy factor of the fuel and Φ determined by following equation [28],

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𝐻

𝑂

𝑆

𝐻

𝛷 = 1.0401 + 0.1728 𝐶 + 0.0432𝐶 + 0.2169𝐶 × (1 ‒ 2.0628 𝐶 )………. (5)

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Where, H, C, O, S are the mass fraction of Hydrogen, Carbon, Oxygen and Sulphur respectively present in the fuel. The exergy rate of the exhaust gas was calculated as summation of its physical and chemical exergy rates as,

𝑐ℎ ………………………………………………….…. (6) 𝐸3 = 𝐸𝑝ℎ + 𝐸 3 3

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The physical exergy of the exhaust gas is calculated by the following equation [24]

[ [

204 205 206

( )]

𝐸𝑝ℎ 3 = (𝑚1 + 𝑚2) 𝐶𝑝,3 𝑇3 ‒ 𝑇0 ‒ 𝑇0𝑙𝑛

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𝑇3 𝑇0

( )]………… (7)

+ 𝑅𝑇0𝑙𝑛

𝑃3 𝑃0

Where, Cp,3, T3 and P3 are the specific heat of exhaust gas at constant pressure, exhaust temperature and exhaust pressure respectively. The specific heat of exhaust gas is determined by the following equation, 𝑚𝑤,3 × 𝐶𝑝,𝑤 × (𝑇𝑐,𝑂 ‒ 𝑇𝑐,𝐼)

𝐶𝑝,3 = (𝑚

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mw,3

) × (𝑇𝑒𝑥,𝐼 ‒ 𝑇𝑒𝑥,𝑂)

1 + 𝑚2

……………………………….……….. (8)

Cp,w are the water flow rate in calorimeter and specific heat of water. Tc,I and are the water temperature at inlet and outlet of calorimeter. Tex,I and Tex,O are the exhaust

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Where,

and

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Tc,O

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gas temperature at inlet and outlet of the calorimeter.

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The chemical exergy of the exhaust gas is calculated by the equation no.9,

∑ ∑ ………………(9) 𝐸𝑐ℎ 3 = (𝑚1 + 𝑚2)𝑛( 𝑖𝑥𝑖𝜀𝑖 + 𝑅𝑇0 𝑖𝑥𝑖ln (𝑥𝑖))

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n is the mole number per kilogram of exhaust gas. xi is the molar fraction of each of the component, єi is the standard chemical exergy of each component and R is the gas constant

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(0.008314 kJ/mol K).

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The standard molar percentage of different species in dead state is given as,

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(N2)=75.67; (O2) (kr)=0.000076.[29].

213

218 219 220 221

Where,

=

20.34;

H2O(g)=3.03;

(CO2)=0.03;

(Ar)=0.00052;

(Ne)=0.0018;

During combustion process, the above standard molar percentage of different species changes and new combinations appear. Hence, the molar percentages of different species generated through the combustion process were recalculated for the exhaust gas to determine its chemical exergy. Then,

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the chemical exergy of the exhaust gas was calculated by using the standard chemical exergy of the exhaust gas components [27] .

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The exergy rate of the shaft work produced is equal to the energy rate of the work done,

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i.e; 𝐸

4=

2𝜋𝜔𝜏 …………………………………………………..………….(10) 60

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Where, ω and τ are angular velocity and torque of the engine.

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The exergy loss rate to the cooling water is calculated by equation 11,

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[

( )] ….………………..………………(11)

𝐸5 = 𝑚5𝐶𝑝,𝑤 𝑇5,𝑜𝑢𝑡 ‒ 𝑇5,𝑖𝑛 ‒ 𝑇0𝑙𝑛

𝑇5,𝑜𝑢𝑡 𝑇5,𝑖𝑛

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Where, 𝑚5 , Cp,w, T5,out and T5,in are the cooling water mass flow rate, specific heat of water, outlet temperature and inlet temperature of the cooling water of the engine.

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The exergy loss to the environment is calculated as following,

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(

)

𝑇0

𝐸6 = 𝑄6 1 ‒ 𝑇

………………………………………………. (12)

6

Where, 𝑄 is the heat loss to the environment and T6 is the average engine body temperature. The 6 heat loss to the environment is calculated by equating the total input energy to the total output energy as presented by equation 13,

𝑄6 = 𝑚1𝐶𝑝1(𝑇1 ‒ 𝑇0) + 𝑚2𝑞𝐿𝐻𝑉 ‒ [(𝑚1 + 𝑚2)𝐶𝑝,3(𝑇3 ‒ 𝑇0)] ‒ 𝑚5𝐶𝑝,5(𝑇5,𝑜𝑢𝑡 ‒ 𝑇5,𝑖𝑛) ‒

2𝜋𝜔𝜏 60

…………….….(13)

The exergetic efficiency of the engine is calculated by equation no 14 as shown below,

𝜑=𝐸

𝐸4

1 + 𝐸2

……………………………………….………..(14)

Exergy analysis of any heat engine can be used to improve it’s efficiency and sustainability. A sustainable use of the energy resources along with their efficient use is vital for sustainable development. Exergy analysis establishes a relation between efficiency and sustainability of a system. Sustainability of any such system can be measured by means of sustainability index of the machine, which can be derived from the following equation [30],

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1

𝑆𝐼 = 1 ‒ 𝜑

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…………………………………………………..(15)

Entropy generation rate (𝑆𝑔𝑒𝑛) is expressed as the ratio of exergy destruction rate and dead state temperature. It is an important parameter in analyzing the losses of an engine as it allows more exact calculation of the losses than the traditional approach involving loss correlations. It is a relatively new tool for improving the overall efficiency of the engine [31] .

𝐸𝑑𝑒𝑠

𝑆𝑔𝑒𝑛 =

250

𝑇0

………………….………………… (16)

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3 Measured Data Uncertainty Analysis

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There are a number of physical and operational parameters that causes some degree of uncertainty

254

during working with an instrument. Hence, an uncertainty analysis with respect to the repeatability

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and precision of the experimentation is of prime importance for assuring accuracy. To this end, a

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comprehensive error estimation was conducted by means of a combined uncertainty analysis for the

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performance parameters. The analysis was conducted on basis of the root mean square method,

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where the total uncertainty U of a quantity Q has been estimated, depending on the independent

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variables x1, x2,.,xn (i.e., Q=f [x1,x2,.,xn]) having individual errors ∆x1, ∆x2,..,∆xnas given by Eq. -17

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[18]. The percentages of uncertainty of the performance parameters are shown in Table 3.

U  (

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Computed Performance Parameter

BP (Brake Power)

Fuel flow

U U U X 1 ) 2  ( X 2 ) 2  ...  ( X n ) 2 ……………………….(17) X 1 X 2 X n

Measured Variables

Instrument Involved in Measurement

% Uncertainty of the Measuring Instrument [5]

Load, RPM

Load Sensor, Load Indicator, Speed measuring Unit.

0.2, 0.1, 1.0.

SFC (Liquid Fuel) Fuel Measuring Unit, Fuel Flow Transmitter

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Total % Uncertainty of the Computed Parameters

(0.2) 2  (0.1) 2  (1.0) 2

1.02.

(0.065) 2  (1.5) 2

1.501

0.065, 1.5.

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Calculation

Table-3: Uncertainty analysis of Performance Parameters

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The measuring range and accuracy of the AVL Digas 444 emission analyser is given in Table-4 [18] Instrument Measuring Range Accuracy Carbon Monoxide (CO) 0-10% Vol < 0.6% vol: ±0.03% vol> 0.6% vol: ± 5% Carbon Dioxide(CO2) 0-20% Vol < 10% vol: ±0.5% vol> 10% vol: ± 5% vol Hydrocarbon (HC) 0-20000 ppm Vol < 200 ppm vol: ± 10 ppm vol> 200 ppm vol: ± 5% NOx 0-5000 ppm Vol. < 500 ppm vol: ± 50 ppm vol . ≥ 500 ppm vol: ± 10% Oxygen (O2) 0-22% Vol < 2% vol: ±0.1% vol ≥ 2% vol: ± 5% vol Table: -4: Accuracy of Emission Measuring instruments.

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4. Result and Discussion.

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The effect of the fuel blends at the tested load conditions has been analyzed with reference to the

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performance, combustion and emission characteristics of the test engine.

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4.1 Performance Analysis.

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4.1.1 Brake Thermal Efficiency

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Fig-3: Variation of Brake thermal efficiency with load for different fuel blends

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The brake thermal efficiency of an engine is a function of the thermal input from the fuel. It is used to evaluate how well an engine converts the heat from a fuel to mechanical energy. The variation of brake thermal efficiency (ƞbth) for the tested blends at different load conditions is shown in Fig-3. It can be seen that at 20% load condition, all fuel samples produced very low efficiency. This is due to very low BMEP, which degrades the combustibility of fuel inside combustion chamber. At 40% load, it can be seen that Diesel produced comparatively higher ƞbth as compared to other fuel samples. This may be because the whole system being calibrated for diesel operation , can perform

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optimum with Diesel. It can also be seen from the figure that the brake thermal efficiency of the engine is increased with increase in ethanol percentage upto 15% (v/v). This consistent improvement is due to superior combustion quality with increase in ethanol percentage. As the ethanol percentage increases, is brings down the viscosity of the biodiesel, allowing better injection of the blends. The improved injection facilitates better atomization of the charge. This allows the large and small chain hydrocarbons of diesel and biodiesel to dissociate quickly into smaller elements during combustion. This dissociation also allows the fuel bound oxygen of ethanol and biodiesel to participate in the combustion process and improve the combustion behavior of the charge. Traces of this improved combustion are also evident in the cylinder pressure development curves shown in Fig 5a-f. However, Fig-3 also displays a decrease in brake thermal efficiency with D30E20B50 blend. This decrease is due to a number of reasons, such as over dilution of the blends due to higher percentage of ethanol that not only affects its injection behavior, causing lower injection pressure and decreased atomization, but also decreases the calorific value low enough to prohibit effecting and significant conversion of energy to work. Side by side, the cooling effect caused by such large percentage of ethanol may also reduce the combustion chamber temperature during vaporization, thus decreasing combustibility

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4.1.2 Brake Specific Energy Consumption

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Fig: 4: Variation of Brake specific energy consumption with load for different fuel blends

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Brake specific energy consumption indicates the efficiency of the engine with which the input energy content of the fuel is utilized by the engine during combustion and it has been expressed in Eq-18.

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( )=

𝐵𝑆𝐸𝐶

𝑘𝐽 𝑘𝑊ℎ

(𝑚𝐷)(𝐿𝐻𝑉𝐷) 𝐵𝑃

+

(𝑚𝐵)(𝐿𝐻𝑉𝐵) 𝐵𝑃

+

(𝑚𝐸)(𝐿𝐻𝑉𝐸) 𝐵𝑃

………… (18)

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Where, mD is the mass flow rate of Diesel, mB is the mass flow rate of biodiesel, mE is the mass flow rate of ethanol. The lower calorific value of Diesel, biodiesel and ethanol are represented as LHVD, LHVB and LHVE respectably.

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The BSEC of the engine at tested load conditions for the tested blends is shown in Fig-4. Apart from 40% load condition, the increment of ethanol upto 15% showed a decreasing trend of BSEC. This consistent decrease in BSEC indicates better utilization of the chemical energy of the fuel into useful work. The improved combustion behavior of engine subjected to the tested ethanol enriched blends ensures higher percentage of fuel energy release during oxidation. This observation is consistent with improved cylinder pressure generation, regarded as an index of combustion quality and shown in Fig-5a-f. It is also worth noticing that a consistent increase in BSEC is observed with D30E20B50 blend. This increase indicated higher fuel energy consumption for the said bend in comparison to its predecessor fuel combinations. This increase indicates a depletion in combustion quality as indicated by reduced brake thermal efficiency (Fig-3) and reduced cylinder pressure (Fig5a-f).

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4.2 Combustion Analysis

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4.2.1 Cylinder Pressure

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Fig: 5a-f: Variation of Cylinder Pressure for different blends at tested load conditions

Variation of in cylinder pressure for the tested blends at different load conditions has been shown in Fig-5a-f. The in-cylinder pressure is an index of quality of combustion inside combustion chamber. As the fuel oxidizes, it releases heat and subsequently the gas mixture expands inside the combustion chamber and increases the cylinder pressure. The cylinder pressure profiles shown in Fig-5a-f clearly show an increase in cylinder pressure, except D45E5B50 blend at all tested load conditions. The maximum cylinder pressure is achieved invariably with D35E15B50 blend, indicating better quality of combustion. The viscosity plays an important part in combustibility of the fuel as proper injection would be achieved from an injector designed for diesel injection if the fuel viscosity is not very high as compared to diesel. Biodiesel has a higher viscosity compared to diesel. Hence, the injection deteriorates causing improper atomization and incomplete combustion. Addition of ethanol reduces the viscosity of the blend and eventually the fuel viscosity reaches an ideal condition with 15% ethanol. As a result, D35E15B50 blend delivers good atomization characteristics and improved combustion, which subsequently leads to high in-cylinder pressure profile. Further, the presence of 15% of ethanol in the blend also improves the combustibility, by means of the high oxygen content of the blends due to inclusion of oxygen rich ethanol and biodiesel. It is also seen that as the ethanol percentage in the blend goes beyond 15%, the incylinder pressure of the engine gradually decreases. This indicates that the combustion quality degrades with percentage increase in ethanol beyond 15% concentration. This retardation in combustion is due to excessive decrease in fuel viscosity, which hampers injection and atomization. With the higher percentage of ethanol, the cumulative viscosity decreases beyond proper injection range and hinder the injection process. As a result, the combustion drags towards incompleteness and causes subsequent decrease in cylinder pressure.

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4.2.1 Heat Release Rate

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Fig:6a-f: Variation of heat release rate for different blends at tested load conditions.

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The heat release rate from the engine is calculated from the measured cylinder pressure data by using the First Laws of Thermodynamics applicable to the closed part of engine cycle expressed as per Eq. 19 [32]

𝑅𝑂𝐻𝑅 =

𝑑𝑄 𝑑𝜃

=

𝐶𝑉 𝑅

(

𝑑𝑉

𝑑𝑝

= 𝑝.𝑑𝜃 + 𝑉.𝑑𝜃 ‒

)+𝑃

𝑃𝑉 𝑑𝑀 𝑚 𝑑𝜃

𝑑𝑉 ……………….. (19) 𝑑𝜃

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The heat release rate from the engine for the tested blends at the 6 tested load conditions is shown in Fig.6a to f. It can be seen from the graphs that the heat release rate for the engine decreased for D45E5B50 blend, which may be due to retardation of combustion process caused by higher viscosity of the blend owing to high percentage of biodiesel in it. It is also observed that further increase of ethanol percentage the heat release rate is increased up to ethanol percentage of 15% concentration. The heat released from a fuel depends upon several factors such as the calorific value of the fuel, combustion quality, equivalence ratio etc. The improvement in heat release rate is an indicative of increase in burning rate of the charge inside the combustion chamber. This increase in burning rate may be due to a number of reasons. The decrease in blend viscosity due to ethanol addition improves the injection characteristics of the engine. Side by side it also provides excess oxygen during combustion process to elevate the burn rate. It is further observed that D30E20B50 blend with 20% ethanol again produced marginally reduced heat release rate. This decrease may be due to over thinning of the charge due to higher percentage of ethanol in the charge.

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4.3 Exergy Analysis.

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The exergy analysis of the engine for the tested blends is analysed based on its exergetic efficiency, exergy destruction rate, entropy generation and sustainability index.

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4.3.1 Exergetic Efficiency

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Fig: 7: Variation of exergetic efficiency with load for different fuel blends

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Exergetic efficiency of any heat engine indicates the efficiency of the engine taking the second law of thermodynamics into account. This parameter compares the system to a thermodynamically perfect one operating under the same conditions. Fig-7 presents the effect of different engine loads on the exergy efficiency of the diesel engine fueled with various fuels. It can be seen from the figure that the exergetic efficiency of the engine follows the trend of thermal efficiency shown in Fig-3. Additionally, the exergetic efficiency was found to be lower than the brake thermal efficiency. This may be due to the fact that a portion of the energy supplied to the engine is lost due to thermodynamic irreversibilities such as severe chemical reactions, friction, heat transfer through walls etc [21].

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It is also observed that exergetic efficiency of the engine is increased with the increase in ethanol percentage in the blends upto it 15% (v/v) concentration. This is because of the increment of fuel flow rate owing to consistent decrease in blend calorific value and inverse relation between fuel exergy and exergetic efficiency [33]. The decrease in exergetic efficiency with D30E20B50 blend may be due to degraded combustion quality due to lower viscosity, high latent of vaporization of biodiesel and reduced calorific value of the blend due to higher participation of ethanol.

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4.3.2 Exergy Destruction Rate

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Fig: 8: Variation of exergy destruction rate for different fuel blends at tested load conditions.

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In any actual process, exergy is always destroyed due to the internal irrevesibilites of the system. Thus, the rate of exergy destruction has a profound effect on the efficiency of a system. Fig-8 shows the rate of exergy destruction of the engine at different load conditions for the tested fuels blends. It can be seen from the figure that the exergy destruction rate gradually increases with load

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for all tested fuel samples. This is because, as the load increases, the increased BMEP improves the combustibility inside the combustion chamber and subsequently the heat release and cylinder temperature increases. This increased in-cylinder temperature widens the temperature gradient between the engine and the surrounding and consequently the exergy destruction rate increases.

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It is also observed from the figure that increasing ethanol percentage reduced the exergy destruction rate. The heat absorption during vaporization of ethanol increases with percentile increase in ethanol in the tested blends. As a result, the thermal gradient between the engine and the surrounding decreases with increasing ethanol percentage and it leads to lower exergy destruction rates. However, with D30E20B50 blend, the drop in blend calorific value is such that it increases fuel flow rate to produce necessary brake power at a certain load level. This led to a higher entropy generation (as shown in Fig-9) and exergy destruction due to the massive thermochemical transformation of the fuel and subsequent mixing of the combustion products with intake air [21].

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4.3.3 Entropy Generation

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Fig: 9: Variation of entropy generation for different fuel blends at tested load conditions.

Rate of entropy generation determines the performance of a heat engine. It is an index of irreversibility of an actual system. Fig -9 shows the variation of entropy generation rate of the engine at different load conditions for the tested fuel combinations. It is seen from the graph that the entropy generation rate has a similar trend with that of exergy destruction rate as they are linearly correlated to each other. Hence, the exergy destruction can be interpreted as the reason for entropy generation during the combustion process.

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4.4 Emission Analysis

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4.4.1 NOx Emission

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Fig-10: Variation of NOx emission with load for different loads.

The increase in NOX emission from a CI engine is primarily due to increased combustion temperature and equivalence ratio [18]. The NOX emission from the engine is shown in Fig. 10. It can be seen from the graphs that the NOX emission from the engine increases with load irrespective of the fuel used. This is because, increase in load increases the BMEP of the engine and thus improves the combustion quality of the engine. Improvement in combustion quality gives rise to higher cylinder temperature, which in terns increases NOX emissions. It is also seen that other than 120% load condition, the fuel blends showed reduced or similar NOX emission as compared to base Diesel operation. Despite of improvement in combustion quality, this decrement in NOX emission is due to high latent heat of vaporization of ethanol. Ethanol has a latent heat of vaporization of 846kJ/kg [18] .Hence during vaporization ethanol absorbs significant amount of heat from the combustion chamber and drops its temperature. As a result, the temperature of combustion chamber is reduced and this restricts the NOX formation mechanism. At 120% load condition, the cooling effect of ethanol in over powered by the high cylinder temperature and consequently a sharp spike in NOX emission is observed.

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4.4.2 Unburned HC Emission

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Fig-11: Variation of unburned HC emission with load for different loads.

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Unburned hydrocarbon emission from a CI engine is due to improper injection causing incompleteness in combustion process. The HC emission from the engine is shown in Fig.11. It is seen that the HC emission from the engine decreased with increasing ethanol percentage upto 15% concentration. Addition of ethanol improves the injection characteristics of the engine and adds up extra oxygen inside the combustion chamber as it dissociates. As a result, the breaking of long chain hydrocarbons of Diesel and biodiesel happens effectively and their oxidation happens at a higher oxygen concentration. This ensures burning of higher percentage of hydrocarbons and reduces the unburned hydrocarbon emission. It is also observed that the HC emission again increases with D30E20B50 blend. As the percentage of ethanol increases, it reduces the cumulative cetane number of the blend due to its very low cetane content. As a result, the ignition delay period may increase causing accumulation of charge inside combustion chamber. Consequently, a fuel rich environment is created inside the combustion chamber and this leads to incompleteness in combustion that results in increased HC emission. It may also be noticed that the unburned hydrocarbon emission from the engine is relatively higher at 120% load condition as compared to full load condition for some of the fuel samples. This may be because, 120% load is overload condition for engine. At such high load, the constant RPM test engine draws higher amount of fuel to produce more power in order to cope up with the load. However, the oxygen supply remains more or less same, as the air intake is consistent irrespective of the load. Hence, because of the limited oxygen available inside the combustion chamber, all the injected fuel does not burn properly and produces unburned HC and CO. Hence, higher amount of HC and CO is detected in the exhaust. Fig.12 is also in accordance with this finding.

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4.4.3 CO Emission

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The CO emission from the engine is shown in Fig.12. It is evident from the Figure that the CO emission from the engine decreases with the increasing load. This is because, as the load increases, the BMEP of the engine increases along with it. This increase in the BMEP is conducive to better combustion inside the combustion chamber. As the CO emission is a result of the incomplete combustion of the fuel carbon, any improvement in the quality of combustion reduces the CO emission.

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It is also observed from the graph that addition of ethanol gradually decreased the CO emission from the engine. This is because during oxidation ethanol releases OH. These 'OH' radical is conducive for oxidation of CO into CO2 [34]. Therefore, much of the CO is oxidized to CO2 and subsequently the CO emission gets reduced. The lowest CO emission is invariably shown by D35E15B50 blend. This indicates that ethanol-biodiesel percentage ratio in this blend is most suitable in terms of oxidation of CO into CO2. As the load increased, the higher BMEP improved the combustibility of the charge, which in terms aided the oxidative effect of D35E1B50 blend and produced lowest CO emission. However, the blend D30E20B50 showed an increasing trend in CO emission from the engine. This sudden increase is due to retardation of combustion due to over thinning of the charge, that caused partially incomplete combustion, giving rise to unburned carbon and CO emission.

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Fig-12: Variation of unburned CO emission with load for different loads.

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4.5 Comparison of Energy Efficiency and Exergy Efficiency

Fig-13 (a-f): Comparison of exergy efficiency and brake thermal efficiency in increasing ethanol percentage.

Fig. 13(a-f) shows a comparison between exergy efficiency and brake thermal efficiency in correlation with increasing ethanol and engine load percentages. It is noticed from the figures that the exergy efficiency of the engine for different ethanol percentages in consistently lower than the brake thermal efficiency of the engine. This reduction in exergy efficiency is due to the fact that, a portion of the input energy is lost due to thermodynamic irreversibilities, that occur during processes such as fuel burning, intensive chemical reactions, severe heat transfers, mechanical

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friction etc. [27]. Further, it is also noticed that both exergy efficiency and brake thermal efficiency of the engine was highest for D35E15B50 blend with 15% ethanol in it. This non-trivial increase in exergy efficiency and brake thermal efficiency indicates a substantial improvement in chemical energy utilization of the fuel and conversion of that energy into effective work.

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4.6 NOX-Sustainability Index-BSEC Tradeoff

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Fig-14: Tradeoff between NOx emission- sustainability index and BSEC at 20% & 40% load.

Fig-15: Tradeoff between NOx emission- sustainability index and BSEC at 60% & 80% load.

Exergy has high potential in improving the efficiency of a system by minimizing the detrimental losses to the environment and maximizing efficient use of the resource. Therefore, the exergy analysis can be applied to adjust the sustainability of the thermal system [35]. Keeping this in mind with the impending contingencies of the diesel engine, a comprehensive NOx-Sustainability IndexBSFC tradeoff analysis has been undertaken in the present study to visualize the effectiveness of the Diesel-ethanol-biodiesel blends and have been summarized in Fig-14, 15 and 16 corresponding to 20%-40% load, 60%-80% load and 100%-120% load conditions. The effect of a particular fuel sample at a particular load on NOx emission, Sustainability Index and BSFC is represented in a symbolic manner. For example, effect of D30E20B50 blend at 20% load condition is represented by ‘20 E20’. The tradeoff between NOx-Sustainability Index-BSFC footprints at 20% and 40% load points is sown in Fig.-14. Evaluation of tradeoff signatures at 20% point reveals a marginally higher BSEC for all tested blends, with base diesel operation being the one with highest BSEC. It also reveals that the BSEC of the engine gradually decreases with ethanol addition at the expanse of higher NOx emission. However, the D35E15B50 blend (represented here as E15) registered a marginally higher sustainability index as compared to rest of the fuel samples. Further increase in ethanol percentage in E20 blend showed a marginal decrease in sustainability with appreciably lower NOx emission. At 40% load conditions, it is seen that the base diesel operation has the lowest sustainability in terms of energy with NOx emission on the higher side of the graph. The BSEC of the engine at this

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load point is low for all fuel samples except D45E5B50 blend. Similar to 20% load condition, here also it is seen that the increasing the ethanol percentage upto 10% sowed a marginal increase in BSEC and NOx emission but at 15% ethanol concentration, both BSEC and NOx emission decreased. Additionally, the sustainability index also higher for E15 blend as compared to E5 or E10 blends. It also worth mentioning that E20 blend again showed marginally lower sustainability, similar BSEC and marginally lower NOx emission as compared to E15 blend. The tradeoff comparison at 60% load condition shown in Fig.15 reveals very high BSEC and appreciably low sustainability, along with significantly high NOx emissions for base diesel and E20 blends. E5 blend also showed minimal NOx emission at the expanse of moderate rise in BSEC and very low sustainability. The advantage of ethanol at this load condition is witnessed with E10 and E15 blends. Although, E15 showed slightly higher NOx emission then E10 blend, but the BSEC and sustainability index associated with E15 was found to be much better than that of E10, indicating a higher acceptability. The same trend was also observed with 80% load condition, with E15 blends showing the highest sustainability and lowest BSEC but at a penalty of very high NOx emission. Fig.16 maps the tradeoff front achieved at high load conditions, i.e. at 100% and 120% load conditions. The study of the contour plot reveals both base diesel operation and E20 blends shares similar BSEC footprints, which were far lower than other D-E-B blend operations. Along with that, they also portrayed appreciably low sustainability index at both of the load conditions. E5 blend at both the load conditions demonstrated a lower BSEC, noticeably higher sustainability index and a drastic reduction in NOx emission as compared to base diesel operation. As the ethanol percentage is increased from 5%, it is found that both NOx emission and sustainability index increased along with marginal decrease in BSEC. The best tradeoff characteristics was shown by E15 blend, as it brought the tradeoff to a low BSEC, lower NOx emission and reasonably high sustainability index.

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Fig-16: Tradeoff between NOx emission- sustainability index and BSEC at 100% & 120% load.

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5. Conclusion

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In course of the present study, an experimental investigation was carried out on an existing diesel engine to explore the effect of ethanol and biodiesel on performance, combustion exergy and emission characteristics of the engine. The ethanol percentage was steadily elevated with an incremental factor of 5 %, culminating into 4 blends with maximum ethanol percentage of 20%. The biodiesel percentage was fixed at 50% and subsequently, diesel percentage was decreased steadily with ethanol percentile share. The effects of increasing ethanol percentage ware compared to baseline diesel operation. Based on the experimental results, following observations were made,

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

The cylinder pressure of the engine initially decreased with ethanol inclusion, but with 10% and 15 ethanol addition is proportionally increased. However, further increase in ethanol percentage again caused a decrease in cylinder pressure. Since cylinder pressure is a direct index of combustion quality, hence it is clear that very low (i.e. 5% concentration by volume) and very high concentration (i.e. above 15% concentration by volume) retards the combustion process. It is also concluded that D35E15B50 blend with 15% ethanol and 50% biodiesel is the best blends among the tested blends in terms of combustion quality. Further, the high heat release rate for the above-mentioned blend also certified the improvement in combustion quality. The results of improvement in combustion are observed in terms of improved brake thermal efficiency and reduced brake specific energy consumption of the engine for D35E15B50 blend. Further, the higher degree of completeness in combustion also ensured lower HC emission for the D35E15B50 blend. However, the actual advantage of ethanol addition was observed, as the

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blend was also able to reduce NOx emission along with HC emission, although in general the emission of HC and NOx remain inversely proportionate to each other. The exergetic efficiency of the engine for base diesel operation was poor as compared to the DE-B blends. Increase in ethanol percentage gradually increased the exergetic efficiency of the blends and eventually D35E15B50 blend produced maximum exergetic efficieny. It is also observed that the exergy destruction rate and entropy generation rate of the above mentioned blend is lowest among the tested blends. This indicates a better usability of the chemical energy of the fuel s compared to base diesel operation. Sustainable development requires efficient use of sustainable, clean and affordable energy resources. Exergy analysis is a powerful tool to indicate the sustainability of a system. Keeping this in mind, a tradeoff study was conducted with sustainability index to investigate the sustainability of the tested fuel blends in the light of two major performance and emission characteristics, i.e. BSEC and NOx emission respectively. It was revealed in the study that D35E15B50 blend is the best suited among the tested blends that can enable a normal CI engine to effectively use the chemical energy of the fuel by minimizing the engine losses with reduced NOx emission.

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Based on the findings of the present study, it is understood that the combustion characteristics of the Diesel-biodiesel fuel blends can be significantly improved by the inclusion of limited amounts of ethanol. In general the outcomes of the study identifies the potential of ethanol for achieving a more cost-effective and eco-friendly combustion process. Further the exergy analysis established the fact that ethanol addition in Diesel-biodiesel blends upto 15% concentration by volume can reduce the losses and improve the energy utilization in the engine.

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