effect of fuel oxygen on the energetic and exergetic efficiency of a compression ignition engine...

lable at ScienceDirect

Energy xxx (2014) 1e9

Contents lists avai

Energy

journal homepage: www.elsevier .com/locate/energy

Effect of fuel oxygen on the energetic and exergetic efficiency of acompression ignition engine fuelled separately with palm and karanjabiodiesels

Jibanananda Jena*, Rahul Dev MisraNational Institute of Technology, Mechanical Engineering Department, Silchar, Assam 788010, India

a r t i c l e i n f o

Article history:Received 25 September 2013Received in revised form18 February 2014Accepted 20 February 2014Available online xxx

Keywords:Energy analysisExergy analysisEfficiencyIrreversibilityBiodieselfuel oxygen

* Corresponding author.E-mail addresses: [email protected] (J. Jena)

D. Misra).

http://dx.doi.org/10.1016/j.energy.2014.02.0790360-5442/� 2014 Elsevier Ltd. All rights reserved.

Please cite this article in press as: Jena J, Mengine fuelled separately with palm and ka

a b s t r a c t

Exergy analysis of any thermodynamic system can take care of the limitations of energy analysis such asirreversible losses, their magnitude and the source of thermodynamic inefficiencies apart from energylosses. In the present study, both the analyses along with heat release analysis are conducted on a naturalaspirated diesel engine fuelled separately with palm biodiesel (PB), karanja biodiesel (KB), and petro-diesel (PD) using the experimental data. Since the engine performs best at about 85% loading condition,the energetic and exergetic performance parameters of the engine are evaluated at 85% loading conditionfor each type of fuel. The aim of the study is to determine the effect of fuel oxygen on energy and exergyefficiencies of a CI (compression ignition) engine. Various exergy losses, exergy destruction and theirratios associated with the heat transfer through cooling water, radiation, exhaust gas, friction, and someuncounted exergy destruction are investigated. Apart from exergy loss due to heat transfer; the un-counted exergy destruction (due to combustion) also plays a major role in the system inefficiency. Basedon the comparative assessment of the obtained results, it is concluded that a better combustion with lessirreversibility is possible with the increase in O2 content in the fuel.

� 2014 Elsevier Ltd. All rights reserved.

1. Introduction

The compression ignition (CI) engine is the most preferredprime mover in many applications, owing to its reliability com-bined with excellent fuel efficiency. In general, CI engines aredesigned to run with petroleum fuel (fossil fuel). Combustion offossil fuel is the major source of CO2 and GHGS emissions to theatmosphere, which resulting severe environmental problem likeGlobal warming and unnatural climate change. In this regard,biofuels may be considered as one of the alternative fuel optionsprovide a partial solution to both these problems, by replacingfossil fuel use and thereby reducing CO2 concentration and GHGemissions. The fuel properties of these biofuels are similar to pet-rodiesel in most ways and hence may be used with little or noengine modification. So exploitation of biofuel efficiently in the CIengine is highly required. In order to analyse engine performanceand to evaluate quantitatively the inefficiencies associated withvarious processes; second law analysis is a better option. Second

, [email protected] (R.

isra RD, Effect of fuel oxygenranja biodiesels, Energy (2014

law analysis deals with the key word e ‘‘exergy” that explains thepotential of the system to produce useful work. Unlike energy,exergy can be destroyed, which is a result of some phenomena suchas combustion, friction, mixing, throttling etc [1]. The exergydestruction is a source for insufficient use from fuel exergy toproduce useful mechanical work in an IC engine. The reduction ofexergy destruction (irreversibility) can lead to better engine per-formance by more efficient exploitation of fuel [2]. Flynn et al. [3]explained a new observation in IC engine studies. They developeda computer model for second law analysis of a turbocharged dieselengine under transient condition. It was reported that combustionirreversibility was the important factor for system inefficiency andtransient in-cylinder irreversibilities were different from steady-state. Alasfour [4] applied an energy and exergy analysis to an SIengine operating at steady-state, to evaluate the use of a butanolegasoline blend as fuel and found that energetic efficiency was about28% of the fuel input energy. Canakci and Hosoz [5] presented acomparative study of energy and exergy analyses for a 4-cylinderturbocharged diesel engine fuelled with various biodiesels andpetrodiesel. Caliskan et al. [6] applied exergy analysis to a JohnDeere 4045T diesel engine run with no. 2 diesel fuel, Soybean oilMethyl Easter and High-Oleic Soybean oil Methyl Easter at

on the energetic and exergetic efficiency of a compression ignition), http://dx.doi.org/10.1016/j.energy.2014.02.079

Delta:1_given name

Delta:1_surname

mailto:[email protected]

mailto:[email protected]

www.sciencedirect.com/science/journal/03605442

http://www.elsevier.com/locate/energy

http://dx.doi.org/10.1016/j.energy.2014.02.079



J. Jena, R.D. Misra / Energy xxx (2014) 1e92

1400 rpm. They found the thermal efficiencies between 39.93% and41.31% and exergetic efficiencies are between 37.46% and 38.48%with no statistically significant difference. Zhang et al. [7] havestudied the effect of ambient oxygen concentration on biodieseland diesel spray combustion under simulated compression ignitionengine conditions for a constant-volume chamber. They haveconcluded that, 18% ambient O2 condition worked better for bio-diesel than petrodiesel in reducing soot particle. With 12% ambientO2 condition, diesel combustion was significantly degraded. How-ever, both fuels experienced low temperature combustion at 10% ofambient O2. Thus, biodiesel could be able to achieve the desiredlower soot production under a moderate oxygen level with highercombustion efficiency. Mohsen et al. [8] have investigated on theexergy recovery from a turbocharged OM314 DIMLER diesel engineby varying the engine speeds (1200, 1400, 1600, 1800 and2000 rpm) and torques (20, 40, 60, 80 and 100 N m). They haveused a double pipe heat exchanger in the exhaust of the enginewith counter current flow. They have found that the recoveredexergy was increased with the increase of load and engine speed.Further, they reported that using recovered exergy, bsfc wasdecreased by approximately 10%. Misra et al. [9] have appliedexergy analysis method to a diesel engine run with petrodiesel andpalm biodiesel. They reported that the Exergetic efficiency of theengine running with PB was approximately 2.12% higher thanpetrodiesel. Ismail and Mehta [10] discussed a method of esti-mating the availability destructions and exergetic efficiencies ofcombustion for different fuels, viz. hydrogen, hydrocarbons, alco-hols and biodiesel surrogates. It was reported that availabilitydestruction is greater for heavier hydrocarbon fuels and oxygen-ated fuels with higher oxygen fraction. The maximum and mini-mum exergetic efficiency was found for hydrogen and acetylenefuel, respectively. Further they have found that availabilitydestruction increases with exhaust gas recirculation (EGR) anddecreases with oxygen enrichment of the supplied air. Ozkan et al.[11] have investigated the effect of the dwell time of multiple in-jection events on the energetic and exergetic efficiencies of a CIengine. The test engine was run using three different injectionstrategies. Using the experimental data of the engine, the heatrelease rate, combustion temperature, heat balance, thermal effi-ciency, and exergetic efficiency were calculated. No significantdifference in the energetic efficiency, exergetic efficiency and alsoon various engine irreversibilities was obtained. But they reportedthat the nitrogen oxides (NOx) emissions were decreased by 7.4%via implementing appropriate pre-injection mass and injectionadvance. Thus from the available literature it may be inferred thatno significant contribution has been reported regarding the influ-ence of the irreversibility components on the performance of CI

Fig. 1. Schematic diagram of experim

Please cite this article in press as: Jena J, Misra RD, Effect of fuel oxygenengine fuelled separately with palm and karanja biodiesels, Energy (201

engine operated on biofuels with respect to the O2 content by thefuel. Once this influence is established, more intensive researchmay be possible in finding the effective biofuel mix for better en-gine performance.

Since all the biodiesels developed so far are having oxygencontent in the range of 8e12% by wt, therefore any biodiesel(s)may be suitable for this purpose. Although edible biodiesels areeasily available but their demand for human consumptionrestricted them to be used as biodiesels in many countries. In thisregard non-edible biodiesels are having advantage over theiredible counterpart. However, organized plantations of non-edibleoil bearing trees are very limited, even though a number of suchoils already have been established. Considering these aspects,one established non-edible biodiesel e karanja biodiesel (KB)and one established edible biodiesel e palm biodiesel (PB) wereconsidered in this work. Further, the conventional fuel for theengine, i.e. the petrodiesel (PD) was also considered for effectivecomparison.

In this study, an attempt has beenmade to study the influence offuel oxygen on various irreversibility components, exergetic effi-ciency and energetic efficiency of a single cylinder, 4-stroke dieselengine operated separately on petrodiesel (PD), palm biodiesel (PB)and karanja biodiesel (KB). The considered fuels are tested in theengine at various loading conditions, viz., at 0%, 25%, 50%, 75%, 85%,90% and 100%. It has been found that the engine performance isbest at about 85% loading condition, which is also supported byavailable literature [12]. The reaction equations, energy rate bal-ance and exergy rate balance for the engine are determined byusing the experimental data of 85% loading condition. The energyand exergy analyses are then investigated by solving the combus-tion reaction equation using (CO2, CO, H2O, H2, O2, NO, N2, andunburned HC) as combustion products. Subsequently, various en-ergetic and exergetic performance parameters of the engine areevaluated for each fuel and compared with each other.

2. Material and method

2.1. Experimental setup

The engine setup consists of a single cylinder, 4-stroke watercooled diesel engine with compression ratio of 17.5 and 7 bhp at1500 rpm. The uses of this type of engine are found in rural/agri-cultural sector for running the irrigation pump-sets, small capacityelectrical generators etc. A Labview based engine performanceanalysis software package “EnginesoftLV,” Apex Innovations,1995 [13] is used for online performance evaluation.

ental setup of engine test rig.

on the energetic and exergetic efficiency of a compression ignition4), http://dx.doi.org/10.1016/j.energy.2014.02.079

J. Jena, R.D. Misra / Energy xxx (2014) 1e9 3

A schematic diagram is shown in Fig. 1 with the locations ofvarious sensors. The specifications of different components ofexperimental engine setup are tabulated in Table 1.

The experiments were carried out in the aforesaid experimentalsetup with twin objectives of engine performance and emissionsmeasurements for the considered fuel samples. For testing aparticular fuel the engine is to run on petrodiesel for about 15 minand then changed over to the particular fuel by the change overcocks and allowed to run for about 7e10 min so that the fuel iscompletely changed over without any traces of petrodiesel. Allobservations and calculated operational performance parameters(through “EnginesoftLV”) are recorded. Various engine parameterslike speed, load, cooling water inlet temperature to engine as wellas calorimeter, cooling water outlet temperatures from engine andcalorimeter, exhaust gas temperatures, air flow rate, fuel flow rate,and water flow rates through engine and calorimeter are recordedvia online “EnginesoftLV” software. The observed values of variousparameters are utilized to calculate engine torque, brake power,frictional power, indicated power, BTE, BSFC, A/F ratio, heat utili-zation in brake power development, exhaust loss, heat loss tocooling water and heat loss by radiation. Engine emissions likecarbon dioxide, carbon monoxide, nitrogen oxides, unburned hy-drocarbon and oxygen aremeasured by the help of AVL-444 five gasanalyser and smoke was measured using AVL-437 smoke meter.The tests with a particular fuel are carried out three times in thesame day without varying in any condition, for checking therepeatability of the tests. The results of the three repetitions wereaveraged. The various properties of the tested fuels are tabulated inTable 2.

2.2. Determining the reaction equations

Energy and exergy analyses of the test engine for each testoperation have been performed after solving the combustion re-action equation for the considered operation. In this study, theconsidered emission products are CO2, CO, H2O, H2, O2, NO, N2, andunburned HC. Some of the engine emission values are presented inTable 3. The amount of solid carbon present in the exhaust gas issufficiently low (<0.5 percent of fuel mass) and omitted from theanalysis NOX mainly consists of nitric oxide [2,14]. Considering this,the combustion reaction can be written explicitly as

CaHbOg þ AS

BðO2 þ 3:76N2Þ ¼ NPðaHCþ bO2 þ cCOþ dCO2

þ eNOþ fN2 þ gH2 þ rH2OÞ(1)

where B is the equivalence ratio, Np is the total no of moles ofexhaust products, and the coefficients a, b, c, d, e, f, g, and r are themole fractions of the respective components. In all test operations,the fuel chemical formulae, fuel and air flow rates and the con-centrations of CO2, CO, O2 in percentage and NO and unburned HCin ppm are known from the experiments. Thus, the coefficients a, b,c, d, and e are known. The four unknowns in Eq. (1) are:Np, f, g and r.Four equations are obtained using the atomic balance of four ele-ments (C, H, N and O).

The energy and exergy analyses of the engine are performedusing the coefficients of various exhaust products considered inEq. (1).

2.3. Energy analysis

The following assumptions are made for the energy analysis ofthe system: the engine runs at steady-state, the combustion air and


exhaust gas each forms ideal gas mixtures, potential and kineticenergy effects of the combustion air, fuel stream and exhaust gasare ignored [15]. Thewhole engine including the dynamometer andcalorimeter is selected as the control volume. The mass and energybalance for an open system control volume can be expressed,respectively as,X

_min ¼X

_mout (2)

_Q � _W ¼X

_mouthout �X

_minhin (3)

where the subscripts in and out represent inlet and exit states,respectively, _Q is the heat rate, _W is the work output, _m is the massflow rate, and h is the specific enthalpy. Net work of the controlvolume, i.e. the brake power, is calculated with experimental datausing Eq. (4).

_Wbrake ¼ uT (4)

where u is angular velocity and T is torque.The fuel energy input rate to the engine is calculated using lower

heating value and mass flow rate ð _mfuelÞ of the fuels as follows:

_Efuel ¼ _mfueljLHVj (5)

The heat carried away by the exhaust gases expressed as a rateof energy flow is calculated by using Eq. (6) [16].

_Qexhaust ¼ _mwCpwðTw2 � Tw1ÞðTe2 � Te3Þ

ðTe1 � T0Þ (6)

where _mw is the mass flow rate of calorimeter cooling water (kg/s), Cpw is the specific heat of calorimeter cooling water (kJ/kg K),Tw1 is the cooling water inlet temperature (�C), Tw2 is the coolingwater outlet temperature (�C), Te1 is the exhaust gas temperatureat the engine (�C), Te2 is the exhaust gas temperature at inletto calorimeter (�C), Te3 is the exhaust gas temperature atoutlet from the calorimeter (�C), and T0 is the ambient air tem-perature (�C).

Heat loss from control volume is consisting of heat loss tocooling water and heat loss due to radiation. The heat transfer byradiation to the environment occurs mainly from hot externalsurfaces (crankcase and cylinder wall) of the engine( _Q radiation ¼ crankcase radiation þ cylinder wall radiation). As theamount of heat carried by lubricating oil comparatively very small,it is neglected in this study. In the present work frictional lossesinclude the pumping loss and blow down loss which gets reflectedin the net work output. Thus, heat loss from the engine controlvolume can be written as,

_Q loss ¼ _Qcoolingwater þ _Q radiation (7)

The heat loss due to engine cooling water can be evaluated as,

_Qcoolingwater ¼ _mcwcpwðTcw2 � Tcw1Þ (8)

where _mcw is the mass flow rate of engine cooling water (kg/s), Tcw2is the cooling water outlet temperature at engine (�C), Tcw1 is thecooling water inlet temperature at the engine (�C).

The heat loss from the engine through radiation as shown inFig. 2 can be evaluated as,

_Q radiation ¼ _Efuel � _Wbrake � _Qexhaust � _Qcoolingwater (9)

The work loss due to friction is calculated as,


Table 1Specifications of experimental engine setup.

Engine Make Kirloskar, Type 1 cylinder, 4-S Dieselengine, water cooled, power 5.2 kW at1500 rpm, stroke 110 mm, bore 87.5 mm661 cc, CR 17.5

Dynamometer Eddy current, water cooled, with loadingunit (Make: SAJ, Pune)

Propeller shaft With universal jointsAir box M.S. fabricated with orifice meter and

manometerFuel tanks Two tanks 7.5 L capacity with glass fuel

metering columnCalorimeter Type pipe in pipePiezo sensor Range 5000 PSI, with low noise cable (Make:

PCB Piezotronics, USA)Crank angle sensor Resolution 1 deg, Speed 5500 rpm with

TDC pulse. (Make: Kubler, Germany)Data acquisition device NI USB-6210, 16-bit, 250 kS/s (Make:

National Instruments, USA)Piezo powering unit Make-Cuadra, Model AX-409.Digital milivoltmeter Range 0e200 mV, panel mountedTemperature sensor Type RTD, PT100 and K type ThermocoupleTemperature transmitter Type RTD PT100, range 0e100 �C,

output 4e20 mA.Type Thermocouple, range 0e1200 �C,Output 4e20 mA

Load indicator Digital, range 0e50 kg, Supply 230VACLoad sensor Load cell, strain gauge type, range 0e50 kgFuel flow transmitter DP transmitter, range 0e500 mm WCAir flow transmitter Pressure transmitter, range (e) 250 mm WCSoftware “EnginesoftLV” Engine performance

analysis softwareRotameter Engine cooling 40e400 LPH;

Calorimeter 25e250 LPHPump Type MonoblockOverall dimensions W 2000 � D 2500 � H 1500 mm

Table 3Engine emission data at 85% loading condition.

Fuel CO2(%) O2(%) CO (%) HC (ppm) NOX (ppm)

PD 6.95 10.92 0.055 30.55 1120.46PB 6.04 11.06 0.046 22.64 1362.56KB 6.67 10.99 0.037 26.40 1237.27


_W friction ¼ _W indicated � _Wbrake (10)

The brake specific fuel consumption is evaluated using,

BSFC ¼ _mf_Wbrake

3600 (11)

where _mf is the mass flow rate of fuel and _Wbrake is the brake work.Thermal efficiency of the control volume is calculated as the

ratio of the net work output to the fuel energy input rate from

h ¼_Wbrake_Efuel

(12)

where _Efuel is the fuel energy input rate to the engine.

Table 2Some properties of the tested fuels.

Properties Petrodiesel Palm biodiesel Karanjabiodiesel

Carbon (% mass)a 86.96 76.33 77.96Hydrogen (%mass)a 13.03 12.40 13.07Oxygen (% mass)a e 11.25 8.97C/H ratio 6.673 6.155 5.964Typical formula C14H25 C18.07H34.93O2 C23.15H46.25O2

Lower heating valueb (kJ/kg) 42,410 38,610 36,975Average molecular weightb 193.35 284.3 257.62Cetane numberb

(ASTM D613)48 55.8 47.5

Kinematic viscosityb

(at 40 �C; mm2/s)2.83 4.64 5.02

Density (kg/m3)b 835 865 877

a Measured at Chemical Laboratory IIT, Mumbai.b Measured at Mechanical Engineering Department NIT, Silchar.


2.4. Exergy analysis

Exergy is the maximum theoretical work which can be obtainedwhen a system of interest interacts with a reference environmentto equilibrium [17]. The exergy destructions and losses in theprocesses and components of a thermal system can be revealed bythe exergy analysis of the system. In this study, the exergy analysisis performed considering the reference environment temperature(T0) of 298.15 K and pressure (p0) of 1 atm. The specific flow exergyof a fluid stream can be found by summing thermo-mechanical andchemical exergies, i.e.,

e x ¼ e xTM þ e xCH (13)

The thermo-mechanical exergy can be defined as

e xTM ¼�h� h0

�� T0ðs� s0Þ (14)

where h and s signify the specific enthalpy and entropy of the fluid,respectively, whereas h0 ands0 stand for the corresponding valueswhen the fluid comes to equilibrium with the reference environ-ment. The specific chemical exergies of liquid fuels can be deter-mined by using Eq. (15) [18].

exCHF ¼�1:0401þ 0:1728

hcþ 0:0432

oc

þ 0:2169Sc

�1� 2:0628

hc

��jLHVj (15)

where h, c, o and s are the mass fractions of H, C, O and S, respec-tively. In this study, the chemical exergies of the fuels are calculatedusing Eq. (15) along with values of H, C, O and S presented inTable 2. In this work, it is assumed that the reactants (air and fuel)enter the engine at the reference state. As such, the thermo-mechanical exergy of fuel and air and the chemical exergy of airare considered to be zero. Thus, the input exergy, which is the fuelexergy, includes only chemical exergy of the fuel, which can bedescribed as,

_E xin ¼ _mfuel � exCHF (16)

The thermo-mechanical exergy of the exhaust gas having ncomponents at the temperature T and pressure p, and can be ob-tained as follows [19]:

Cooling water outAirFuel

Cooling water in

Exhaust gases

Brake power

Lubricating oil in

Lubricating oil outCrankcase and wall radiation

Fig. 2. Energy balance diagram of the system.


Table 4Definition of environment [18].


e xTMexhaust ¼(Xn

ai

�hiðTÞ � hiðT0Þ � T0

�s0ðTÞ � s0ðT0Þ

Reference component Mole fraction (%)

N2 75.6700O2 20.3500CO2 0.03450H2O 3.03000CO 0.00070SO2 0.00020H2 0.00005Others 0.91455

Fig. 3. Exergy balance diagram of the system.

i¼1

� Rinpp0

�)(17)

where ai is the coefficient of the ith component in the reactionequation, s0 is the absolute entropy at the standard pressure (kJ/kmol K), and R is the universal gas constant. T0 and p0 are thetemperature and pressure at environmental state. Using exhaustgas temperatures the values of hi and s0ðTÞ can be found from idealgas property tabular data of Moran and Shapiro [18].

Chemical exergy of the exhaust gases is calculated usingEq. (18) [20].

e xCHexhaust ¼ RT0Xni¼1

ai ln

yiyei

!(18)

where yi is the molar ratio of the ith component in the exhaust gasand yei is the molar ratio of the ith component in the referenceenvironment, given under the definition of the environment inTable 4. Mole fractions of the exhaust gases were calculatedbalancing real combustion equations by means of emissionmeasurement.

The chemical exergy for the substances, which are not present inthe environment (e.g. fuel and NO) can be evaluated by consideringan idealized reaction of the substance with other substances forwhich the chemical exergies are known [19].

The steady-state exergy rate balance equation can be written as,

Xj

1�T0

Tj

!_Qj � _Wcv þ

Xin

_mfiexfi�Xout

_mfoexfo � _Exdest ¼ 0

(19)

where Tj indicates the absolute temperature at the location on theboundary where the heat transfer occurs. The terms ofPin

_mfiexfi andPout

_mfoexfo show the rate of exergy entering and

leaving the control volume accompanying the fuel stream,respectively. _E xdest represents the rate of exergy destroyed in thecontrol volume due to irreversibilities.

_E xQ ¼X�

1� T0T

�_Q|fflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflffl}

Exergy flow

accompanying heat

_E xW ¼ _WBrake ¼ _Wcv|ffl{zffl}Exergy flow

accompanying work

_E xf ¼ _W friction|fflfflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflfflffl}Frictional irreversibility

_E xin ¼ _mfuel � exCHF|fflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl{zfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflfflffl}Fuel exergy

As explained in Fig. 3, for an engine having a fixed power output(brake power or brake exergy), any reduction either in its exergylosses or in its exergy destructions or in both leads towardsreduction in fuel consumption, i.e. reduction in fuel exergy,resulting in higher exergetic efficiency. Since there is a possibility ofextracting some useful work from the exhaust gas and coolingwater, the exergy associated with the exhaust gases and coolingwater are treated here as exergy loss but not exergy destruction,although no special application has been considered to utilize thesame. So the sources of various exergy losses are identified asexergy losses due to heat transfer through cooling water toambient, through exhaust gases to ambient. The sources of various


exergy destructions are associated with heat transfer through ra-diation, through mechanical friction, and some uncounted exergydestruction (mainly due to combustion and fluid flow irreversibil-ities). The total fuel exergy supplied to the engine is distributed intothe exergy associated with brake power and sum of all aboveexergy losses and exergy destructions. Thus, the unaccountedexergy destruction can be calculated as,

_E xðuncountedÞ ¼ _E xin � _Wbrake ��

_E xcw þ _E xrad þ _E xexh þ _E xf�

(20)

Exergetic efficiency can be found as

hII ¼_E xw_E xin

(21)

The entropy generated in the system can be obtained by usingGouyeStodola theorem as,

_E xdest ¼ T0 _Sgen (22)

When rate of exergy destruction and rate of exergy loss ofvarious sources are compared with the total fuel exergy, it gives theexergy destruction ratio ðyD ¼ _E xdest= _E xinÞ and exergy loss ratioðyL ¼ _E xloss= _E xinÞ, which provides thermodynamic measures ofthe system inefficiency for that respective source.

2.5. Heat release rate analysis

The heat release rate of the tested fuels PD, PB and KB areanalysed at 85% loading condition on the basis of recorded cylinderpressure data. The combustion chamber is considered as the con-trol volume for conducting the heat release rate analysis. Thefollowing assumptions are made for the heat release analysis: idealgas behaviour of the fuel-air mixture inside the engine cylinder,uniform composition of the charge from the closing of intake valveto opening of exhaust valve, and the heat energy liberated duringcombustion is the heat added to the cylinder. Applying first law ofthermodynamics to the control volume, the energy balance equa-tions can be written as,

dUdt

¼ _Qc � _Wc (23)

and


Cyl

inde

r P

ress

ure

(bar

)

Crank Angle (deg)

Fig. 4. Variation of cylinder pressure against crank angle of the tested fuels.


mCvdTdt

¼ _Qc � Pdvdt

(24)

where _Qc is the summation of the heat rate generated due tocombustion and the heat rate transferred across the cylinder wall._Wc is the rate of work done by the system boundary in a closedsystem cycle.

Replacing the time variable with the crank angle, ‘q’ and usingthe ideal gas assumption the heat release rate can be found as [21],

_Qc ¼ g

g� 1PdVdq

þ 1g� 1

VdPdq

(25)

where g ¼ Cp/Cv is the specific heat ratio of the gas mixture in thecylinder. The value of g for petrodiesel varies in between 1.3 and1.35. p is the instantaneous cylinder pressure (Pa) and V is theinstantaneous cylinder volume (m3).

2.6. Error analysis

Since the magnitude of error is always uncertain, therefore, it isperhaps better to use the experimental uncertainty. Generally anumber of equipments are integrated in an experimental setup toachieve a common result. Each of the equipment has its own degreeof uncertainty in measurement. The cumulative uncertainty of theoverall results of the experiment is calculated by using thefollowing relationship [22,23].

cEM ¼�ðcFCÞ2 þ ðcLSÞ2 þ

�crpm

�2 þ ðcBTEÞ2 þ ðcCWTÞ2

þðcCOÞ2 þ ðcHCÞ2 þ ðcNOxÞ2 þ ðcSMÞ2 þ ðcPSÞ2�1=2

(26)

where c is the uncertainty.The uncertainties of the components used in the experimental

setup of the present work, taken from the manufacturer’s cata-logue, are given in Table 5. Thus, the total uncertainty of the resultsobtained in the present work is �2.64%.

3. Result and discussions

The experiments were conducted on a single cylinder naturalaspirated CI engine, at various loading condition without changingthe operating condition. It has been found that the engine perfor-mance is best at about 85% loading condition, which is also sup-ported by available literature [10]. Therefore, the heat release andcombustion analyses, the energy analysis, and the exergy analysishave been carried out at this condition.

Table 5Uncertainties of the components used in the experiment.

Sl.No. Components %-Age uncertainties

1 Fuel consumption measurements (FC) �0.52 Load sensor (LS) �0.013 rpm pickup �0.64 BTE �0.155 CWT �0.16 CO �0.037 HC �0.018 NOx �0.39 Smoke meter (SM) �110 Pressure sensor (PS) �0.1


3.1. Heat release and combustion analyses

The variation of in-cylinder pressure of PD, PB and KB is pre-sented in Fig. 4. The peak pressure of the tested fuels is found to be72.17 bar for PD, 75.24 bar for KB and 75.60 bar for PB. It is observedthat the peak pressure in case of PB and KB is higher than that of PD.Again comparing the same between the tested biodiesels, PB hasgot the higher peak pressure than KB. This is because of higherCetane No (CN) of PB reducing the ignition delay and higher oxygenpercentage of PB, promoting better combustion and leading tohigher peak pressure. This is in agreement with published litera-ture [24]. Combustion in CI engine takes place primarily in twophases: premixed combustion and diffusion combustion. The fac-tors that govern the relative amounts of these two phases ofcombustion are engine load, injection timing, and CN of the fuel.The combustion in CI engine with biodiesel is greatly influenced bythe CN and oxygen content of the fuel [24]. Increased CN reducesthe ignition delay resulting in less intense premixed combustionphase. However, the higher heat release rate of petrodiesel asshown in Fig. 5 is due to the higher accumulation of fuel during therelatively longer delay period. Because of shorter ignition delay, themaximum heat release rate occurs early for biodiesels than whencompared with petrodiesel. As a result of improved combustionduring the main combustion phase due to the higher oxygen con-tent of the fuel. The heat release rate for the tested biodiesels is lessduring the late combustion phase than petrodiesel. This is becausethe amount of oxygen left over after the main combustion phasehelps to burn the fuel in the late combustion phase and ensures thecomplete combustion of the fuel. This is attributable for higherexergetic efficiency and lower uncounted exergy destruction.

Fig. 5. Variation of heat release rate against crank angle of the tested fuels.


Table 6Energy analysis results of the fuels tested in the engine.

Fuel Fuel energyrate (kW)

Cooling water heat lossrate (kW)

Radiation heat lossrate (kW)

Exhaust heat lossrate (kW)

Indicated workoutput (kW)

Brake poweroutput (kW)

Energy thermalefficiency (%)

PD 17.17 5.175 1.39 6.04 7.256 4.576 26.65PB 16.33 5.41 1.529 4.815 6.845 4.576 28.02KB 16.42 5.24 1.501 5.103 7.044 4.576 27.86

0

5

10

15

20

Fuelenergy

Coolingwater heat

loss

Exhaustloss

Radiationheat loss

Ene

rgy

valu

es(k

W) PD

PB

KB

Fig. 6. Energy values of tested fuels.


3.2. Energy analysis

The results of energy analysis in terms of input fuel energy, heatloss, exhaust loss and brake power for the tested fuels (PD, PB andKB) are presented in Table 6 and Fig. 6. The fuel energy required fora brake power output of 4.57 kW with PD, PB and KB are 17.17 kW,16.33 kW and 16.42 kW, respectively. The fuel energy required forPB and KB are less than petrodiesel by 4.9% and 4.37%, respectively,because fuel energy is proportional to LHV and the LHV of bio-diesels are slightly less than the LHV of petrodiesel. This is inagreement with the published literature [25,29].

It is observed in Fig. 6 that the exhaust energy loss rate for PB &KB is less than petrodiesel by 5.69% and 4.1%, respectively. This isbecause of the exit temperature of exhaust gas is less for biodieselsthan that of petrodiesel. The energy loss due to exhaust gas isproportional to the temperature difference between exhaust gastemperatures and surrounding. This is confirmed from the lowerconcentration of CO and unburned HC presence in the exhaustgases for biodiesels as shown in Table 3. When the engine isoperating with biodiesels (PB and KB) it is observed that heatrejection rate in the form of cooling water and radiation to theatmosphere is higher than PD by 4.25% and 3%, respectively. It canbe argued that as biodiesels yield less CO and unburned HC in theexhaust, it indicates promotion of better combustion with bio-diesels; as a result a higher rate of fuel energy transfer is takingplace in the form of work and heat from the engine. This is also

Bra

ke t

herm

al e

ffic

ienc

y(%

)

Fig. 7. Brake thermal efficiency values of the tested fuels.


confirmed from the early peak heat release rate for biodiesels asshown in Fig. 10. Again, comparing the heat loss rate and exhaustenergy loss rate between PB and KB it is found that the exhaust lossrate with PB is less than that with KB, where as the heat loss rateassociated with radiation and cooling water for PB is higher thanthat with KB. This inferred that combustion is better for PB than KB.This may be due to a higher percentage of oxygen present in PB,which helps to accelerate the combustion in the main combustionphase.

Fig. 7 shows the variation of brake thermal efficiency of thetested fuels. The engine operated with PB and KB yield higher BTEthan petrodiesel by 1.37% and 1.21%, respectively. As BTE is the ratioof power output to input fuel energy, therefore for the sameamount of power output, biodiesels provided slightly less energy tothe engine than petrodiesel due to slightly less in LHV.

The variations of brake specific fuel consumption of the testedfuels are shown in Fig. 8. The result shows that for PB and KB theBSFC rate are higher than PD by 4.5% and 10% approximately. This isbecause the LHV of biodiesels are less than LHV of PD; as a resultthe difference in LHV is compensated by higher rate of fuel con-sumption in case of biodiesels for a same brake power output. Themass flow rate of fuel to the engine in case of biodiesels is higherthan petrodiesel may also be due to higher viscosity and densitiesin comparison to that of petrodiesel.

3.3. Exergy analysis

The results of the exergy analysis for the same brake power of4.576 kW in terms of fuel exergy rate, exergy loss rates, exergyratios, exergy destruction rates and exergy destruction ratios arepresented in Table 7. Fuel exergy rates of PD, PB and KB are 18.3 kW,17.54 kW, and 17.63 kW, respectively. Since the specific exergy ofeach fuel type is proportional to the lower heating values, the fuelexergy values follow similar trend as that of their energy counter-parts. The fuel exergy rates for PB and KB are lower than PD by4.15% and 3.66%, respectively.

It is seen that the exergy loss ratio due to cooling water for PD,PB and KB are 1.55%, 1.94% and 1.87%, respectively. This exergy lossis proportional to the temperature difference between the com-bustion temperature and the cooling water temperature. Thecombustion temperature of PB is the highest followed by KB andthen PD. This may be due to better combustion in case of oxygen

BSF

C ( k

g/kW

hr)

Fig. 8. Brake specific fuel consumption values of the tested fuels.


Fig. 9. Exergetic efficiency values of the tested fuels.

Table 7Exergy analysis results of the fuels tested in the engine.

Exergy rate Exergy values (kW) Exergy loss ratios (yL) (%) Exergy destruction ratios (yD) (%)

PD PB KB PD PB KB PD PB KB

_E xin 18.30 17.54 17.63 e e e e e e_E xcw 0.283 0.339 0.329 1.55 1.93 1.87 e e e_E xexh 4.670 4.201 4.226 25.52 23.95 23.97 e e e_E xfriction 2.03 2.159 2.14 e e e 11.09 12.31 12.14_E xrad 3.263 3.461 3.414 e e e 17.83 19.73 19.36_E xuncounted 3.480 2.8 2.943 e e e 19.01 15.96 16.69


richer fuels (biodiesels). As a result, the exergy loss to enginecooling water for PB is the highest followed by KB and then PD.

The exergy loss ratio through exhaust is the minimum for PBfollowed by KB and then PD. The calculations for PB, KB and PDshow that 23.95%, 23.97% and 25.52% of fuel exergy is lost from theengine through exhaust gases. The exhaust exergy loss ratios sus-tain a similar trend with exhaust energy loss to ambient. The trendis similar with available literature [5,26,27].

The exergy destruction ratio due to frictional irreversibility forPD, PB and KB are 11.09%, 12.31% and 12.14%, respectively. Thefriction loss is more for biodiesels, because for a same brake powerlarger fuel mass flow rate to the engine as the viscosity and densityboth are higher than PD. The uncounted exergy destruction ratio(mainly due to combustion irreversibility) is 19.01%, 15.96% and16.69% for PD, PB and KB, respectively. This exergy destruction forbiodiesels is slightly less than petrodiesel because biodiesels are ofoxygen richer fuel promoting better combustion in the combustionphase and also provided better lubricity in the diffusion phase dueto higher density and viscosity. The exergy destruction ratiothrough cylinder wall radiation for PD, PB and KB are 17.83%, 19.73%and 19.36%, respectively. These exergy destruction ratios sustain asimilar trend with exergy loss ratios due to cooling water.

The exergetic efficiency of the engine for the tested fuels PD, PBand KB are 25%, 26.01% and 25.95%, respectively, as shown in Fig. 9.The results follow the similar trend with the brake thermal effi-ciency for the same fuels. However, the magnitude of the exergeticefficiencies are lower than brake thermal efficiencies for the same

Fig. 10. Entropy generation rate of the engine for the tested fuels.


fuels, because the engine uses higher amount of fuel exergycompared to fuel energy. These results are in agreement with theavailable literature [28]. Entropy generation rate of the enginerunning with different fuels are presented in Fig. 10. Entropy isgenerated due to system irreversibility. Entropy generation rate isthe maximum for PD, because the exergy destruction rate for it isthe maximum.

4. Conclusions

In the present work, the exergetic performance of a CI enginein terms of irreversibility components and its emissions understeady-state operating condition has been studied without anyengine modifications. The energy and exergy analyses have beencarried out using the experimental performance and emissionsdata for the three selected fuels, namely, petrodiesel (PD) palmbiodiesel (PB) and karanja biodiesel (KB). Based on those, theenergetic and exergetic performance parameters of the engineare evaluated. The petrodiesel is a better quality fuel than bio-diesel, because of its higher net calorific value. That is, the BSFCof biodiesel is higher compared to that of petrodiesel for thesame work output. As expected, the same has been found outthrough the exergy analysis. The magnitudes of the entropygenerations, exergy losses, exergy destructions, exergy loss ra-tios, and exergy destruction ratios have been evaluated. It isfound that uncounted exergy destroyed (mainly due to com-bustion irreversibility) for PB is the minimum, whereas brakethermal efficiency and exergetic efficiency is the maximum.Although petrodiesel is a better quality fuel, the irreversibilitiesproduced with petrodiesel is found to be higher compared to thatwith tested biodiesels. Further, palm biodiesel is found to bebetter than karanja biodiesel in terms of both energetic andexergetic performance. The comparative assessment of the ob-tained results reveals that with the increase in O2 content in thefuel, the combustion on one hand is found to be better and ir-reversibilities get reduced on the other. This interesting resultmay invite more intensive research towards development ofeffective biofuel mix for this kind of CI engine.

Acknowledgements

The authors are very grateful to the reviewers for their valuableand constructive comments, which have been utilized to improvethe quality of the paper.

Nomenclature

BSFC Brake specific fuel consumptionCWT Cooling water temperature_E Energy rate (kW)E x,

Exergy rate (kW)ex Exergy (kJ/kg)e x Exergy (kJ/kmol)



h Specific enthalpy (kJ/kg)LHV Lower heating value_m Mass flow rate (kg/s)N Total no of molesP Pressure (kPa)R General gas constant (kJ/kmol K)s Specific entropy (kJ/kg K)T Temperature (�C or K)T Torque (N m or kN mm)_Q Heat transfer rate (kW)_W Work output (kW)Y Mol fraction (%)

Greek symbolsh Thermal efficiency (%)hII Exergetic efficiency (%)u Angular velocityF Equivalence ratioc Uncertainty

Subscriptscw Cooling waterdest Destructionin Inputout Outputp Productgen Generation

SuperscriptsCH Chemical exergyTM thermo-mechanical

References

[1] Tat ME, Wang PS, Van Gerpen JH, Clemente TE. Exhaust emissions from anengine fuelled with biodiesel from high-oleic soybeans. J Am Oil Chem Soc2007;84:pp.865e869.

[2] Heywood JB. Internal combustion engine fundamentals. New York: McGraw-Hill; 1988.

[3] Flynn PF, Hoag KL, Kamel MM, Primus RJ. A new perspective on diesel engineevaluation based on second law analysis; 1984. SAE paper no.840032.

[4] Alasfour FN. Butanol a single cylinder engine study: availability analysis. ApplTherm Eng 1997;17:537e49.

[5] Canakci M, Hosoz M. Energy and exergy analysis of a diesel engine fuelledwith various biodiesels. Energy Sour Part B 2006;1:379e94.


[6] Caliskan H, Tat ME, Hepbasil A, Van Gerpen JH. Exergy analysis of enginesfulled with biodiesels from high oleic soybeans based on experimental values.Int J Exergy 2010;7:20e36.

[7] Zhang J, Jing W, William LR, Fang T. Effects of ambient oxygen concentrationon biodiesel and diesel spray combustion under simulated engine conditions.Energy 2013;57:722e32.

[8] Mohsen G, Mohammad H, Davood DG, Mofid G, Ali B, Shahi G. Exergy recoveryfrom the exhaust cooling in a DI diesel engine for BSFC reduction purposes.Energy 2014;65:44e51.

[9] Misra RD, Jibanananda Jena, Murthy MS. Energy and exergy analyses of a CIengine fuelled with palm biodiesel based on experimental data. Int J Exergy2013;13:124e39.

[10] Ismail S, Mehta PS. Evaluation of the effects of fuel and combustion-relatedprocesses on exergetic efficiency. Fuel 2011;90:1818e25.

[11] Ozkan M, Ozkan DB, Ozener O, Yılmaz H. Experimental study on energy andexergy analyses of a diesel engine performed with multiple injection strate-gies: effect of pre-injection timing. Appl Therm Eng 2013;53:pp.21e30.

[12] Misra RD, Murthy MS. Straight vegetable oils usage in a compression ignitionengine e a review. Renew Sustain Energy Rev 2010;14:3005e13.

[13] Apex Innovatios Private Limited, Sangli.http://www.apexinnovations.co.in;1995.

[14] Dincer I. Thermodynamics, exergy and environmental impact. Energy Sour2000;22:723e32.

[15] Horlock JH, Winterbone DE. The thermodynamics and gas dynamics of in-ternal combustion engines, vol. II. Oxford: Clarendon Press; 1986.

[16] Kopac M, Kokturk L. Determination of optimum speed of an internal com-bustion engine by exergy analysis. Int J Exergy 2005;2:pp.40e54.

[17] Van Gerpen JH, Shapiro HN. Second law analysis of diesel engine combustion.J Eng Gas Turb Power 1990;112:129e37.

[18] Kotas TJ. The exergy method of thermal plant analysis. Malabar, FL: KriegerPublishing Company; 1995.

[19] Bejan A, Tsatsaronis G, Moran M. Thermal design and optimization. New York:Wiley; 1996.

[20] Moran MJ, Shapiro HN. Fundamentals of engineering thermodynamics. 3rd ed.New York: John Wiley and Sons; 2000.

[21] Tat ME. Cetane number effect on the energetic and exergetic efficiency of adiesel engine fuelled with biodiesel. Fuel Process Technol 2011;92:1311e21.

[22] Holman JP. Experimental methods for engineers. Tata McGraw Hill Publica-tions; 2004.

[23] Devan PK, Mahalakshmi NV. Performance, emission, and combustion char-acteristics of poon oil and its diesel blends in a DI diesel engine. Fuel 2009;88:861e70.

[24] Sahoo PK, Das LM. Combustion analysis of Jatropha, Karanja and Polangabased biodiesel as fuel in a diesel engine. Fuel 2009;88:994e9.

[25] Wallace SJ. Diesel engine energy balance study operating on diesel and bio-diesel fuels. The Department of Mechanical Engineering and the Russ Collegesof Engineering and Technology; 2007. Msc thesis.

[26] Hazar H. Effects of biodiesels on a low heat loss diesel engine. Renew Energy2009;34:pp.1533e1537.

[27] Di Y, Cheunng CS, Huang Z. Comparison of the effect of biodieselediesel andethanolediesel on the gaseous emission of a direct-injection diesel engine.Atmos Environ 2009;43:2721e30.

[28] Sezer I, Bilgin A. Exergy analysis of SI engines. Int J Exergy 2008;5:204e17.[29] Kamolpatara L, Sompong P, Tanet A, Samai J. Impact of using B100 biodiesel in

ship engines. Kasetsart J Nat Sci. 2010;44:1191e200.


http://refhub.elsevier.com/S0360-5442(14)00213-8/sref1










































http://www.apexinnovations.co.in















































effect of fuel oxygen on the energetic and exergetic efficiency of a compression ignition engine...

Documents