Exergetic Analysis of Using Oxygenated Fuels in Spark-Ignition (SI)Engines†

Ismet Sezer,*,‡ Ismail Altin,§ and Atilla Bilgin|

Besikduzu Vocational School, Karadeniz Technical UniVersity, Trabzon 61800, Turkey, Trabzon VocationalSchool, Karadeniz Technical UniVersity, Trabzon 61300, Turkey, and Mechanical Engineering Department,

Karadeniz Technical UniVersity, Trabzon 61080, Turkey

ReceiVed April 15, 2008. ReVised Manuscript ReceiVed July 29, 2008

In this study, the use of the oxygenated fuels in spark-ignition (SI) engines has been investigated by meansof exergy analysis. For this purpose, a two-zone quasi-dimensional thermodynamic cycle model was used.The cycle model contains compression, combustion, and expansion processes. The combustion period issimulated as a turbulent flame propagation process. Intake and exhaust processes are computed by a simpleapproximation method. Principles of the second law of thermodynamics are applied to the cycle model toperform exergy analysis. Exergetic terms, such as exergy transfer with heat, exergy transfer with work,irreversibilities, thermomechanical exergy, fuel chemical exergy, and total exergy, were computed in exergyanalysis. Additionally, distributions of fuel exergy, the energy-based (the first law) efficiency, and the exergetic(the second law) efficiency were calculated. The results showed that the oxygenated fuels are suitable from anexergy point of view because of less entropy production and less heat loss. On the other hand, these fuelscause reduction in work output and an increment in fuel consumption because of their lower calorific valuesand lower stoichiometric air/fuel ratios in comparison to isooctane. Irreversibilities for methanol and ethanolare lower by about 7.44 and 4.29%, respectively, than that of isooctane. Exergy transfer with heat decreasesby about 9.47% for methanol and 6.45% for ethanol in comparison to isooctane. However, exergy transferwith work, i.e., useful work output, decreases by about 7.35 and 3.24% for methanol and ethanol, respectively.The brake-specific fuel consumptions for methanol and ethanol are higher, about 132.2 and 65.5%, respectively,in comparison to isooctane.

1. Introduction

Today’s energy crises and environmental problems haveconcentrated the investigations on alternative fuels for decreasingthe consumption of exhaustible petroleum reserves and mini-mizing the concentration of toxic components.1 Alcohols canbe considered as suitable alternative fuels because they can bemade from renewable resources, such as various grown cropsand even waste products.1,2 Moreover, alcohols reduce theharmful emissions, such as carbon monoxide (CO) and unburnedhydrocarbon (UHC) emissions, by supplying leaner combustionbecause of the oxygen content in their molecular structures.3

Methanol and ethanol are commonly used alcohols as enginefuels or fuel additives because of their fuel properties.4 The fuelproperties of isooctane, methanol, and ethanol are given in Table

1.5-7 Alcohol fuels have simple molecular structures. They burnefficiently and improve combustion efficiency.8 High octanenumbers of methanol and ethanol allow for the use of highercompression ratios and improve thermal efficiency of the engine.Methanol and ethanol have a higher latent heat of vaporizationin comparison to isooctane. This provides more mass into the

† From the Conference on Fuels and Combustion in Engines.* To whom correspondence should be addressed. Telephone: +90-462-

377-29-50. Fax: +90-462-325-55-26. E-mail: isezer@ktu.edu.tr.‡ Besikduzu Vocational School.§ Trabzon Vocational School.| Mechanical Engineering Department.(1) Al-Baghdadi, M. A. R. S. A simulation model for a single cylinder

four-stroke spark ignition engine fueled with alternative fuels. Turk. J. Eng.EnViron. Sci. 2006, 30, 331–350.

(2) Kowalewicz, A.; Wojtyniak, M. Alternative fuels and their applica-tion to combustion engines. Proc. Inst. Mech. Eng., Part D 2005, 219, 103–124.

(3) Ahouissoussi, N. B. C.; Wetzstein, M. E. A comparative cost analysisof biodiesel, compressed natural gas, methanol and diesel for transit bussystem. Resour. Energy Econ. 1997, 20, 1–15.

(4) Sezer, I.; Bilgin, A. Effects of methyl tert-butyl ether addition tobase gasoline on the performance and CO emissions of a spark ignitionengine. Energy Fuels 2008, 22 (2), 1341–1348.

(5) Sezer, I. Experimental investigation of the effects of blendingmethanol and MTBE with regular gasoline on performance and exhaustemissions of SI engines. M.S. Thesis, Karadeniz Technical University,Trabzon, Turkey, 2002 (in Turkish).

(6) Shenghua, L.; Clemente, E. R. C.; Tiegang, H.; Yanjv, W. Study ofspark ignition engine fueled with methanol/gasoline fuel blends. Appl.Therm. Eng. 2007, 27 (11-12), 1904–1910.

(7) Bayraktar, H. Experimental and theoretical investigation of usinggasoline-ethanol blends in spark-ignition engines. Renewable Energy 2005,30, 1733–1747.

(8) http://www.faqs.org/faqs/autos/gasoline-faq/part4/preamble.html.

Table 1. Comparison of Fuel Properties5-7

property isooctane methanol ethanol

chemical formula C8H18 CH3OH C2H5OHmolecular weight (kg/kmol) 114.224 32.04 46.07oxygen percent (wt %) 49.9 34.8density (g cm-3) 700 792 789freezing point at 1 atm (°C) -107.378 -97.778 -80.00boiling temperature at 1 atm (°C) 99.224 64.9 74.4auto-ignition temperature (°C) 257.23 463.889 422.778latent heat of vaporization at 20 °C (kJ/kg) 349 1103 840stoichiometric air/fuel ratio (AFRs) 15.2 6.47 9.0lower heating value of fuel (kJ/kg) 44 300 20 000 26 900LHV of stoichiometric mixture (kJ/L) 3810 3906 3864research octane number (RON) 100 111 108motor octane number (MON) 100 92 92

Energy & Fuels 2009, 23, 1801–1807 1801

10.1021/ef8002608 CCC: $40.75 2009 American Chemical SocietyPublished on Web 08/30/2008

cylinder by cooling the inducted air and increases engine power.9

For these reasons, numerous experimental and theoretical studieshave been performed on the use of oxygenated fuels in internalcombustion engines.7,9-14 However, the investigations devotedto exergy analysis on spark ignition (SI) engines using oxygen-ated fuels are very limited.15,16 Gallo and Milanez15 performedan exergetic analysis for ethanol- and gasoline-fueled SI enginesby using a simulation model. They used a finite rate heat releasemodel for combustion and compared the exergy destructionsduring combustion for ethanol and gasoline for the same engineat various conditions. They found that ethanol gave lessirreversible combustion and higher exergetic efficiency thangasoline even at the same compression ratio. Chavannavar16

completed a parametric study on availability destruction duringcombustion for eight different fuels at combustion conditionsof constant pressure, constant volume, and constant temperature.His work was mainly intended to investigate various parametersaffecting the efficiency of the combustion process. He foundthat increases in the fuel/air equivalence ratio generally resultedin decreases in the destruction of availability. He also foundthat the destruction of availability increases with increasingcomplexity of the molecular structure of the fuel. He concludedthat the results of his study provided significant informationand hinted for design and operation of combustion systems foroptimal use of energy contained in the fuels. As seen in theliterature review, there are very limited numbers of studies onexergy analysis of SI engines using oxygenated fuels and mostof them use a finite heat release combustion model. Burnduration is an input data in some heat release models, and it istaken as a constant for simplicity. Hovewer, burn duration variesdepending upon the fuel type, because different fuels havedifferent combustion characteristics. For this reason, to revealthe effects of using various fuels in a realistic way, combustionis simulated as a turbulent flame propagation process in thisstudy differently from existing studies.15,16

The present study aims to investigate the use of oxygenatedfuels, such as methanol and ethanol, in SI engines via exergyanalysis by using a two-zone quasi-dimensional thermodynamiccycle model by incorporating a turbulent entrainment combus-tion model. The details of the cycle model and exergy analysisare given in the next section.

2. Mathematical Model

2.1. Combustion Simulation. Two zones (i.e., burned andunburned zones) grow in the combustion chamber after the

beginning of combustion. Each zone is assumed to be uniform intemperature and homogeneous in composition. It is also assumedthat a uniform pressure distribution exists throughout the combustionchamber. Combustion is simulated as a turbulent flame propagationprocess, and it is supposed that the flame front develops sphericallythroughout the unburned gases. The governing equations andthe details of the model are found in the literature.17-21 In thecombustion model, laminar flame speed is calculated by usingthe equations developed by Gulder.22 The geometrical features ofthe flame front are computed from a geometric submodel installedon the basics17,18 by using the mathematical relations given in theliterature.23 Further details on the combustion model and itssubmodels can be found in Sezer.24

2.2. Computation of Cycle. As known, SI engine cycles consistof four consecutive processes: intake, compression, expansion, andexhaust. Intake and exhaust processes are computed by using theapproximation method given by Bayraktar and Durgun.21 Compres-sion, combustion, and expansion processes have been computedfrom the governing equations,25 with arrangment of them for eachprocess in a suitable manner. Further details of the cycle modelcan be found in the literature.24,25 The engine performanceparameters, i.e., brake mean effective pressure (bmep) and brake-specific fuel consumption (bsfc), have also been determined fromthe well-known equations.21,24,25

2.3. Exergy Analysis. The second law of thermodynamics hasbeen applied to the above model for the exergetic computations.As known, the second law can be stated by the statement of entropybalance26

∆S)∫ (QT )boundary

+ σ (1)

where, σ is the entropy production because of the irreversibilities.The exergy, i.e., availability, equation for a closed system can

be obtained by using the combination of the first and second lawsof thermodynamics as follows:26,27

A)E+ p0V- T0S (2)

where E ) U + Ekin + Epot, V and S are the total (sum of internal,kinetic, and potential) energy, volume, and entropy of the system,respectively, and p0 and T0 are the fixed pressure and temperatureof the dead state.

Availability is defined as the maximum theoretical work that canbe obtained from a combined system (combination of a system andits reference environment) when the system comes into equilibrium(as thermal, mechanical, and chemical) with the environment.26,27

The maximum available work from a system emerges as the sum

(9) Gao, J.; Jiang, D.; Huang, Z. Spray properties of alternative fuels:A comparative analysis of ethanol-gasoline blends and gasoline. Fuel 2007,86, 1645–1650.

(10) Gautam, M.; Martin, D. W., II. Combustion characteristics of higheralcohol/gasoline blends. Proc. Inst. Mech. Eng., Part A 2000, 214, 497–511.

(11) Abu-Zaid, M.; Badran, O.; Yamin, J. Effect of methanol additionon the performance of spark ignition engines. Energy Fuels 2004, 38, 312–315.

(12) Hsieha, W.-D.; Chenb, R.-H.; Wub, T.-L.; Lin, T.-H. Engineperformance and pollutant emission of an SI engine using ethanol-gasolineblended fuels. Atmos. EnViron. 2002, 36, 403–410.

(13) Donnelly, R. G.; Heywood, J. B.; LoRusso, J.; O’Brien, F.; Reed,T. B.; Tabaczynski, R. J. Methanol as an automotive fuel. MIT EnergyLaboratory Report MIT-EL 76-013, April 1976.

(14) Yacoub, Y.; Bata, R.; Gautam, M. The performance and emissioncharacteristics of C1-C5 alcohol-gasoline blends with matched oxygencontent in a single-cylinder spark ignition engine. Proc. Inst. Mech. Eng.,Part A 1998, 212, 363–379.

(15) Gallo, W. L. R.; Milanez, L. F. Exergetic analysis of ethanol andgasoline fueled engines. SAE Paper 920809, 1992.

(16) Chavannavar, P. S. Parametric examination of the destruction ofavailability due to combustion for a range of conditions and fuels. M.S.Thesis, Texas A&M University, College Station, TX, 2005.

(17) Blizard, N. C.; Keck, J. C. Experimental and theoretical investiga-tion of turbulent burning model for internal combustion engines. SAE Paper740191, 1974.

(18) Keck, J. C. Turbulent flame structure and speed in spark-ignitionengines. International 19th Symposium on Combustion, 1982; pp 1451-1466.

(19) Tabaczynski, R. J.; Ferguson, C. R.; Radhakrishnan, K. A turbulententrainment model for spark-ignition combustion. SAE Paper 770647, 1977.

(20) Tabaczynski, R. J.; Trinker, F. H.; Sahnnon, B. A. S. Furtherrefinement of a turbulent flame propagation model for spark-ignition engines.Combust. Flame 1980, 39, 111–121.

(21) Bayraktar, H.; Durgun, O. Mathematical modeling of spark-ignitionengine cycles. Energy Sources 2003, 25, 651–666.

(22) Gulder, O. Correlations of laminar combustion data for alternativeSI engine fuels. SAE Paper 841000, 1984.

(23) Bilgin, A. Geometric features of the flame propagation process foran SI engine having dual-ignition system. Int. J. Energy Res. 2002, 26,987–1000.

(24) Sezer, I. Application of exergy analysis to spark ignition enginecycle. Ph.D. Thesis, Karadeniz Technical University, Trabzon, Turkey, 2008.

(25) Ferguson, C. R. Internal Combustion Engine Applied Thermo-sciences; John Wiley and Sons: New York, 1985.

(26) Moran, J. M.; Shapiro, H. N. Fundamentals of EngineeringThermodynamics, 3rd ed.; John Wiley and Sons: New York, 1998.

(27) Rakopoulos, C. D.; Giakoumis, E. G. Second-law analyses appliedto internal combustion engines operation. Prog. Energy Combust. Sci. 2006,32, 2–47.

1802 Energy & Fuels, Vol. 23, 2009 Sezer et al.

of two contributions: thermomechanical exergy, Atm, and chemicalexergy, Ach.28 Thermomechanical exergy is defined as the maximumextractable work from the combined system, as the system comesinto thermal and mechanical equilibrium with the environment, andit is given as in eq 3

Atm )E+ p0V- T0S-∑ µ0,imi (3)

where mi and µ0,i are the mass and chemical potential of species i,respectively, calculated at restricted dead state conditions.

At the restricted dead state conditions, the system is in thermaland mechanical equilibrium with the environment and no workpotential exist between the system and environment because oftemperature and pressure differences. However, the system doesnot reach the chemical equilibrium with the environment, becausethe contents of the system are not permitted to mix with theenvironment or enter the chemical reaction with environmentalcomponents. In principle, the difference between the compositionsof the system at the restricted dead state conditions and theenvironment can be used to obtain additional work.28 The maximumwork obtained in this way is called chemical exergy, which can bedetermined as

Ach )∑ mi(µ0,i - µi0) (4)

where µi0 is the chemical potential of species i calculated at the

true dead state conditions.Availability balance for a control volume for any process can

be written as

∆A)A2 -A1 )AQ -AW -Adest (5)

where ∆A is the change of the total system availability for a givenprocess, A2 is the total availability at the end of the process, A1 isthe total availability at the start of the process, AQ is the availabilitytransferred accompanying the heat transfer, AW is the availabilitytransfer because of work, and Adest is destroyed availability byirreversible processes.

Considering fuel chemical exergy, the balance equation for theengine cylinder is written as24,29

dAtot

dθ) (1-

T0

T )dQdθ

- (dWdθ

- p0dVdθ )+ mf

mtot

dxb

dθaf,ch -

dIcomb

dθ(6)

The left-hand side of eq 6 is the rate of change in total exergyof cylinder contents. The first and second terms on the right-handside represent exergy transfer with heat and exergy transfer withwork, respectively. The third term on the right-hand side corre-sponds to the burned fuel exergy, where mf and mtot are the massesof fuel and total cylinder contents and af,ch is fuel chemical exergy,which is calculated by following the equation developed by Kotas30

for liquid fuels.

Af,ch ) af,chmf )QLHV[1.0401+ 0.01728h′c′ + 0.0432

o′c′ +

0.2196s′c′ (1- 2.0628

h′c′ )] (7)

In eq 7, QLHV is the lower heating value of fuel, which is calculatedby using Mendeleyev formula

QLHV ) [33.91c′ + 125.6h′ - 10.89(o′ - s′)- 2.51(9h′ -w′)](8)

The quantities of h′, c′, o′, s′, and w′ in eqs 7 and 8 representmass fractions of elements carbon, hydrogen, oxygen, sulfur, andthe water content in the fuel, respectively.

The last term on the right-hand side of eq 6 illustrates exergydestruction because of combustion. It is calculated as24,29

Icomb ) T0Scomb (9)

where Scomb is the rate of entropy generation because of combustionirreversibilities. It is calculated for two-zone combustion model as

Scomb )d(mbsb)

dθ+

d(musu)

dθ(10)

where mb and mu are burned and unburned masses of the cylindercontents and sb and su are entropy values of the burned and unburnedgases, respectively.

Moreover, total exergy destruction considered here consists ofcombustion and heat transfer irreversibilities as follows:24,29

Itot ) Icomb + IQ (11)

Entropy production sourced from the heat transfer process isgiven in eq 12, and it has been already calculated in eq 6

SQ )Qb

Tb+

Qu

Tu(12)

where Qb and Qu are the rates of heat loss from the burned andunburned zones at temperatures Tb and Tu, respectively.

The efficiency is defined to compare different engine sizeapplications or evaluate effects of various improvements from theperspective of the first- or second-law analysis.27 The first-law (orenergy-based) efficiency is defined as24,25

ηI )energy out (as work)

energy in) W

mfQLHV(13)

where W is indicated work output.Various second-law efficiencies (exergetic efficiency or ef-

fectiveness) have been defined in the literature.26,27 The equationgiven below is used in this study for the second-law efficiency

ηII )exergy out (as work)

exergy in)

AW

mfaf,ch(14)

where AW is exergy transfer with work.

3. Numerical Applications

3.1. Computer Program and Solution Procedure. Acomputer program has been written to perform the numericalcomputations concerning the presented model. Exergetic cal-culations were performed simultaneously depending upon thethermodynamic state of the cylinder content. The resultsobtained have been corrected using the following equations:25

ε1 ) 1- (Vm/V) (44)

ε2 ) 1+ [W/∆(mu)+Qw] (45)

Selecting the values of ε1 and ε2 at the levels of 10-4 was foundconfidently, and the computer program was terminated.

3.2. Validation of the Model. To demonstrate the validationof the present cycle model, the predicted values from the modelare compared to experimental data obtained from the literature.20,31

These comparisons are given in parts a and b of Figure 1 forthe conditions specified on the figure, and engine specificationsare given in Table 2. The burned mass fraction and cylinder

(28) Alkidas, A. C. The application of availability and energy balancesto a diesel engine. J. Eng. Gas Turbines Power 1988, 110, 462–469.

(29) Zhang, S. The second law analysis of a spark ignition engine fueledwith compressed natural gas. M.S. Thesis, University of Windsor, Ontario,Canada, 2002.

(30) Kotas, T. J. The Exergy Method of Thermal Plant Analysis; KriegerPublishing: Malabar, FL, 1995.

(31) Rakopoulos, C. D. Evaluation of a spark ignition engine cycle usingfirst and second law analysis techniques. Energy ConVers. Manage. 1993,34 (12), 1299–1314.

Exergetic Analysis of Oxygenated Fuels Energy & Fuels, Vol. 23, 2009 1803

pressure are selected as comparison parameters. As seen in thefigures, predictions are in good agreement with the experimentaldata. Therefore, it can be said that the presented model has anenough level of confidence for analysis of engine performanceand parametric investigation.

4. Results and Discussion

After the present model has been constructed and controlled,the exergetic computations were performed. The engine speci-fications25 for the parametric investigation devoted to exergyanalysis are given in Table 2.

Parts a-f of Figure 2 show the variations of exergetic termswith respect to crank angle for examined fuels. When usingmethanol and ethanol, a larger amount of fuel is required forsupplying the same amount of energy in the cylinder because

of their lower calorific values and also the lower stoichiometricair/fuel ratios as seen in Table 1. Therefore, equivalence ratiosare organized to supply the same exergy entry into the cylinder.Thus, equivalence ratios are equal to 0.92 for methanol and 0.95for ethanol, while it is 1.0 for isooctane as seen Table 3.

Exergy can be transferred to or from a system accompanyingheat and work, similar to energy, but it can be also destroyedby irreversibilities during a process differently from energy.26

The variations of exergy transfer with heat are shown in Figure2a for three fuels. The negative values in this figure show thatthe direction of exergy transfer is from cylinder contents to thecylinder walls, which means that some amount of the exergy islost. Isooctane has the highest values in magnitude, whilemethanol gives the lowest one. These variations can be attributedto both combustion temperatures and combustion durations ofthe fuels. As seen in Table 3, isooctane has a higher combustiontemperature because of a higher heating value than those ofmethanol and ethanol. Further, isooctane has a longer combus-tion duration compared to methanol and ethanol, which extendsthe contact duration of hot gases with cylinder walls. Both ofthese effects cause an increase in the amount of heat transferredand, thus, the exergy transfer with heat, as expected.

Variations in exergy transfer with work are given in Figure2b. In this figure, the negative part of the variations shows thatthe direction of work transfer is from the piston to the cylindercontents. Coversely, the positive part of the variations showsthat work transfer is from the cylinder contents to the piston,which corresponds to the useful work. Three fuels examinedhave almost the same values during compression (before 0 CAD,i.e., before top dead center). However, isooctane gives themaximum values, and ethanol is better than methanol duringexpansion. These variations in exergy transfer with work canbe explained depending upon the calorific values of fuels. Asmentioned in the Introduction and seen in Table 1, the lowercalorific values of methanol and ethanol cause a decrease inwork output and exergy transfer with work, naturally, formethanol and ethanol. Variations in exergy transfer with workare also suitable for the bmep values given in Table 3.

Figure 2c shows the variations of irreversibilities, which havereached the minimum values for methanol. Ethanol gives highervalues than methanol, and isooctane has the greatest irrevers-ibilities as seen in the figure. The reductions in irreversibilitiesfor oxygenated fuels can be based on the simple molecularstructure of these fuels as explained in the literature.16 Ad-ditionally, the shorter combustion durations for oxygenated fuels,as seen in Table 3, provide another contribution to the reductionof irreversibilities as cited in the literature.33 It is also notedthat both of these positive effects on irreversibilities aredominant over the negative effect of lower combustion tem-peratures obtained with oxygenated fuels.

Figure 2d shows the thermomechanical exergy variations forthe examined fuels. The variations in thermomechanical exergyreflect the combination of exergy transfers accompanying heatand work and also contain irreversibilities, as shown in eq 6.Isooctane has the lower values than methanol and ethanol duringcombustion. The faster burning of oxygenated fuels as cited inthe Introduction causes the rising of temperature and pressurein the cylinder earlier for methanol and ethanol in comparisonto isooctane. For this reason, thermomechanical exergy valuesare a bit higher for methanol and ethanol than that of isooctaneduring combustion. However, methanol and ethanol give very

(32) Heywood, J. B. Internal Combustion Engine Fundamentals;McGraw-Hill: New York, 1988.

(33) Caton, J. A. On the destruction of availability (exergy) due tocombustion processes with specific application to internal combustionengines. Energy 2000, 25, 1097–1117.

Figure 1. Comparison of predicted values with experimental data.

Table 2. Specifications of the Enginesa

specification r Xs

D(mm)

S(mm)

Lcr

(mm)div

(mm)Liv

(mm)

experimental engine I20 9.9 0.0 83 74 122 33 4.4experimental engine II31 7 0.1 76.2 111.125 220 30 4.2parametric study engine25 10 0.0 100 80 160 40 5

a The parameters not present in the references, i.e., div and Liv, aredetermined from the mathematical relations given by Heywood.32

1804 Energy & Fuels, Vol. 23, 2009 Sezer et al.

close but the lower thermomechanical exergy values than thoseof isooctane during expansion. These variations in thermome-chanical exergy values during expansion can be based on theprevious statements of exergy transfers with heat and work andalso irreversibilities. The completion of combustion for oxygen-ated fuels in advance results in lower thermomechanical valuesfor methanol and ethanol toward the end of expansion incomparison to isooctane.

Variations in fuel exergy were shown in Figure 2e. Methanol,ethanol, and isooctane have the same fuel exergy values at thestart of the compression because of a fixed equivalence ratio.Different variations in fuel exergy for each fuel occur duringcombustion because of the changing of burn durations for everyfuel, as seen in Table 3. In other words, the fuel exergy isconsumed more quickly during combustion, as the burn rate ofthe fuel increases. On the other hand, there is no remaining

Figure 2. Variations in (a) exergy transfer with heat, (b) exergy transfer with work, (c) irreversibilities, (d) thermomechanical exergy, (e) fuelexergy, and (f) total exergy for the fuels.

Exergetic Analysis of Oxygenated Fuels Energy & Fuels, Vol. 23, 2009 1805

fuel exergies for the three fuels examined at the end of thecombustion period because of the lean mixtures for oxygenatedfuels and stoichiometric mixture for isooctane.

Variations in total exergy values were given in Figure 2f.Total exergy is a summation of fuel exergy and thermome-chanical exergy, as shown in eq 6. Therefore, it reflects thecharacteristics of both of them. For example, the fuel exergy isconstant before the start of combustion, and, the increase intotal exergy comes from the increase in thermomechanicalexergy. At the end of the expansion stroke, fuel exergy decreasesto zero and the exergy transfer with exhaust comes from thethermomechanical exergy. According to this variation, ethanolgives the lowest total exergy values, while isooctane has thehighest ones during expansion. Exergy transfer with heat reachesthe greatest value for isooctane at the end of the expansion asseen in the figure. Methanol also gives very close but highervalues than ethanol; thus, ethanol has the least exergy transferassociated with exhaust.

Figure 3a shows the distributions of fuel exergy on variousexergetic terms for the different fuels examined. The percentagesof irreversibilities are lower for oxygenated fuels in comparisonto isooctane because of explanations above. Methanol gives theleast amount of irreversibilities among examined fuels. Thedistributions of irreversibilities are about 17.53, 18.16, and18.97% for methanol, ethanol, and isooctane, respectively. Thepercentages of exergy transfer with heat are very close for allexamined fuels, but oxygenated fuels have lower values thanisooctane. The distributions of exergy transfer with heat areabout 7.71, 7.97, and 8.52% for methanol, ethanol, andisooctane, respectively. The proportion of exergy transfer withwork is the maximum for isooctane, while the minimum valueis obtained with methanol, as suitable to exergy variations givenabove. The distributions of exergy transfer with work are about35.61, 37.19, and 38.44% for methanol, ethanol, and isooctane,respectively. The percentages of exergy transfer with exhaustare smaller for methanol and ethanol in comparison to isooctaneas seen in the figure. The distributions of exergy transferassociated with exhaust gases are about 34.15, 33.21, and34.25% for methanol, ethanol, and isooctane, respectively.

Variations of the first-law efficiency (ηI), second-law ef-ficiency (ηII), and brake-specific fuel consumption (bsfc) valueswere given in Figure 3b. The first- and second-law efficienciesare similar in characteristics but different in magnitude as seenin the figure. The first-law efficiency is the maximum forisooctane because of maximum work output. Ethanol has a lowerefficiency value than isooctane, and methanol gives the mini-mum first-law efficiency. Similarly, isooctane has the maximumsecond-law efficiency, while methanol gives the minimum value.Ethanol has a value in between those of methanol and isooctane.On the other hand, the fuels examined have shown interestingvariations in terms of bsfc when considering the variations inefficiencies. Isooctane has given the least amount of bsfc, whilemethanol has the greatest one. The increments in bsfc are morethan twice for methanol and more than one and half times forethanol when compared to isooctane. Additionally, it is notedhere that methanol and ethanol have higher research octanenumbers than isooctane as seen in Table 1, which allows for

selection of the higher compression ratios for them. Hence, itis possible to improve engine performance for oxygenates byselecting higher compression ratios as suggested in theliterature.8,15

5. Conclusions

In this study, the use of oxygenated fuels in SI engines hasbeen evaluated by using a two-zone quasi-dimensional cyclemodel via exergy analysis. The following conclusions can bedrawn from the present study: (1) The quasi-dimensional modelof the engine cycle is very suitable for parametric studies, andcoupling of this model with exergy analysis is very useful inassessment of engine operation. (2) Methanol and ethanol aresuitable from the perspective of exergy analysis because of theless entropy production and the lower heat and exhaust losses.The values of irreversibilities decrease about 7.44 and 4.29%for methanol and ethanol, respectively, in comparison toisooctane. Exergy transfer with heat decreases about 9.47% formethanol and 6.45% for ethanol in comparison to isooctane.Additionally, exergy transfer with exhaust decreases in mag-nitude about 0.37 and 3.07% for ethanol and methanol,respectively, in comparison to isooctane. (3) The oxygenatedfuels examined result in a decrease in the work output becauseof their lower heating values in comparison to isooctane. Thus,exergy transfer with work decreases in magnitude about 7.35%for methanol and 3.24% for ethanol in comparison to isooctane.(4) Methanol and ethanol give the lower first-law efficienciesthan isooctane. The decrements are about 0.43 and 0.05%,respectively, for methanol and ethanol. The second-law ef-ficiencies also decrease about 2.83% for methanol and 1.23%

Table 3. Values Obtained from the Cycle Model for ExaminedFuels

fuels φ

mf

(g) mtot

Qf

(J) Af,ch

Tmax

(K)bmep(bar)

∆θb

(deg)

isooctane 1.0 4.356 × 10-2 0.739 1955 2109 2699 12.47 103methanol 0.92 9.047 × 10-2 0.716 1830 2109 2599 11.52 78ethanol 0.95 6.907 × 10-2 0.733 1894 2109 2651 12.05 69

Figure 3. Distributions of (a) exergetic terms and (b) engine perfor-mance parameters for the fuels.

1806 Energy & Fuels, Vol. 23, 2009 Sezer et al.

for ethanol in comparison to isooctane. However, it is notedthat engine efficiency and performance can be increased byselecting higher compression ratios because of the higher octanequality of oxygenates. (5) The brake-specific fuel consumptionis quite higher for especially methanol and ethanol in comparisonto isooctane. The increments in bsfc are about 132.2% formethanol and 65.5% for ethanol in comparison to isooctane.Considering these extremely higher bsfc values for oxygenatedfuels, the use of methanol and ethanol as fuel additives ingasoline seems more reasonable. Thus, it is possible to decreasethe irreversibilities by improving the combustion properties ofgasoline without increased bsfc.

NomenclatureA ) availability or exergy, JAexh ) exergy transfer associated with exhaust gases, JAf,ch ) fuel chemical exergy, JAQ ) exergy transfer with heat, JAtm ) thermomechanical exergy, JAW ) exergy transfer with work, JAtot ) total exergy, Jbmep ) brake mean effective pressure, barbsfc ) brake-specific fuel consumption, g kW-1 h-1

div ) intake valve diameter, mD ) cylinder diameter, mE ) energy, Jf ) mass residual gas fraction, dimensionlessI ) irreversibilities, JLcr ) connecting rod length, mLiv ) maximum intake valve lift, mm ) mass, kgp ) pressure, barR ) cylinder radius, mRs ) radius of spark plug location from cylinder axis, mQLVH ) lower heating value (or calorific value) of fuel, kJ/kgQw ) total heat transfer to cylinder walls, J

T ) temperature, Kr ) compression ratio, dimensionlessS ) entropy, JU ) internal energy, Jxb ) mass fraction burned, dimensionlessXs ) Rs/R ) spark plug location, dimensionless

AbbreViationsCAD ) crank angle degree

Greek Lettersφ ) fuel/air equivalence ratio, dimensionlessηI ) the first-law efficiency, %ηII ) the second-law efficiency, %θ ) crank angle, CADθs ) spark timing crank angle, CAD∆θb ) burn duration, CAD

Subscripts0 ) reference or dead state conditionsb ) burnedch ) chemicalcr ) connecting rodcomb ) combustiondest ) destructionexh ) exhaustf ) fuelkin ) kineticmax ) maximumpot ) potentials ) sparktm ) thermomechanicaltot ) totalu ) unburnedw ) wall

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Exergetic Analysis of Oxygenated Fuels Energy & Fuels, Vol. 23, 2009 1807