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

Energy & Fuels 2009, 23, 1801–1807

1801

Exergetic Analysis of Using Oxygenated Fuels in Spark-Ignition (SI) Engines† I˙smet Sezer,*,‡ I˙smail Altin,§ and Atilla Bilgin| Besikdu¨zu¨ Vocational School, Karadeniz Technical UniVersity, Trabzon 61800, Turkey, Trabzon Vocational School, 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 means of 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 is simulated as a turbulent flame propagation process. Intake and exhaust processes are computed by a simple approximation method. Principles of the second law of thermodynamics are applied to the cycle model to perform 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 exergy analysis. 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 an exergy point of view because of less entropy production and less heat loss. On the other hand, these fuels cause reduction in work output and an increment in fuel consumption because of their lower calorific values and lower stoichiometric air/fuel ratios in comparison to isooctane. Irreversibilities for methanol and ethanol are lower by about 7.44 and 4.29%, respectively, than that of isooctane. Exergy transfer with heat decreases by about 9.47% for methanol and 6.45% for ethanol in comparison to isooctane. However, exergy transfer with 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 have concentrated the investigations on alternative fuels for decreasing the consumption of exhaustible petroleum reserves and minimizing the concentration of toxic components.1 Alcohols can be considered as suitable alternative fuels because they can be made from renewable resources, such as various grown crops and even waste products.1,2 Moreover, alcohols reduce the harmful emissions, such as carbon monoxide (CO) and unburned hydrocarbon (UHC) emissions, by supplying leaner combustion because of the oxygen content in their molecular structures.3 Methanol and ethanol are commonly used alcohols as engine fuels or fuel additives because of their fuel properties.4 The fuel properties of isooctane, methanol, and ethanol are given in Table † From the Conference on Fuels and Combustion in Engines. * To whom correspondence should be addressed. Telephone: +90-462377-29-50. Fax: +90-462-325-55-26. E-mail: [email protected]. ‡ Besikdu ¨ zu¨ 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 application to combustion engines. Proc. Inst. Mech. Eng., Part D 2005, 219, 103– 124. (3) Ahouissoussi, N. B. C.; Wetzstein, M. E. A comparative cost analysis of biodiesel, compressed natural gas, methanol and diesel for transit bus system. Resour. Energy Econ. 1997, 20, 1–15. (4) Sezer, I˙.; Bilgin, A. Effects of methyl tert-butyl ether addition to base gasoline on the performance and CO emissions of a spark ignition engine. Energy Fuels 2008, 22 (2), 1341–1348.

Table 1. Comparison of Fuel Properties5-7 property chemical formula molecular weight (kg/kmol) oxygen percent (wt %) density (g cm-3) freezing point at 1 atm (°C) boiling temperature at 1 atm (°C) auto-ignition temperature (°C) latent heat of vaporization at 20 °C (kJ/kg) stoichiometric air/fuel ratio (AFRs) lower heating value of fuel (kJ/kg) LHV of stoichiometric mixture (kJ/L) research octane number (RON) motor octane number (MON)

isooctane methanol ethanol C8H18 114.224

CH3OH 32.04 49.9 700 792 -107.378 -97.778 99.224 64.9 257.23 463.889 349 1103 15.2 6.47 44 300 20 000 3810 3906 100 111 100 92

C2H5OH 46.07 34.8 789 -80.00 74.4 422.778 840 9.0 26 900 3864 108 92

1.5-7 Alcohol fuels have simple molecular structures. They burn efficiently and improve combustion efficiency.8 High octane numbers of methanol and ethanol allow for the use of higher compression ratios and improve thermal efficiency of the engine. Methanol and ethanol have a higher latent heat of vaporization in comparison to isooctane. This provides more mass into the (5) Sezer, I˙. Experimental investigation of the effects of blending methanol and MTBE with regular gasoline on performance and exhaust emissions 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 of spark 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 using gasoline-ethanol blends in spark-ignition engines. Renewable Energy 2005, 30, 1733–1747. (8) http://www.faqs.org/faqs/autos/gasoline-faq/part4/preamble.html.

10.1021/ef8002608 CCC: $40.75  2009 American Chemical Society Published on Web 08/30/2008

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cylinder by cooling the inducted air and increases engine power.9 For these reasons, numerous experimental and theoretical studies have been performed on the use of oxygenated fuels in internal combustion engines.7,9-14 However, the investigations devoted to exergy analysis on spark ignition (SI) engines using oxygenated fuels are very limited.15,16 Gallo and Milanez15 performed an exergetic analysis for ethanol- and gasoline-fueled SI engines by using a simulation model. They used a finite rate heat release model for combustion and compared the exergy destructions during combustion for ethanol and gasoline for the same engine at various conditions. They found that ethanol gave less irreversible combustion and higher exergetic efficiency than gasoline even at the same compression ratio. Chavannavar16 completed a parametric study on availability destruction during combustion for eight different fuels at combustion conditions of constant pressure, constant volume, and constant temperature. His work was mainly intended to investigate various parameters affecting the efficiency of the combustion process. He found that increases in the fuel/air equivalence ratio generally resulted in decreases in the destruction of availability. He also found that the destruction of availability increases with increasing complexity of the molecular structure of the fuel. He concluded that the results of his study provided significant information and hinted for design and operation of combustion systems for optimal use of energy contained in the fuels. As seen in the literature review, there are very limited numbers of studies on exergy analysis of SI engines using oxygenated fuels and most of them use a finite heat release combustion model. Burn duration is an input data in some heat release models, and it is taken as a constant for simplicity. Hovewer, burn duration varies depending upon the fuel type, because different fuels have different combustion characteristics. For this reason, to reveal the effects of using various fuels in a realistic way, combustion is simulated as a turbulent flame propagation process in this study differently from existing studies.15,16 The present study aims to investigate the use of oxygenated fuels, such as methanol and ethanol, in SI engines via exergy analysis by using a two-zone quasi-dimensional thermodynamic cycle model by incorporating a turbulent entrainment combustion model. The details of the cycle model and exergy analysis are given in the next section. 2. Mathematical Model 2.1. Combustion Simulation. Two zones (i.e., burned and unburned zones) grow in the combustion chamber after the (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 higher alcohol/gasoline blends. Proc. Inst. Mech. Eng., Part A 2000, 214, 497– 511. (11) Abu-Zaid, M.; Badran, O.; Yamin, J. Effect of methanol addition on 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. Engine performance and pollutant emission of an SI engine using ethanol-gasoline blended 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 Energy Laboratory Report MIT-EL 76-013, April 1976. (14) Yacoub, Y.; Bata, R.; Gautam, M. The performance and emission characteristics of C1-C5 alcohol-gasoline blends with matched oxygen content 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 and gasoline fueled engines. SAE Paper 920809, 1992. (16) Chavannavar, P. S. Parametric examination of the destruction of availability due to combustion for a range of conditions and fuels. M.S. Thesis, Texas A&M University, College Station, TX, 2005.

Sezer et al. beginning of combustion. Each zone is assumed to be uniform in temperature and homogeneous in composition. It is also assumed that a uniform pressure distribution exists throughout the combustion chamber. Combustion is simulated as a turbulent flame propagation process, and it is supposed that the flame front develops spherically throughout the unburned gases. The governing equations and the details of the model are found in the literature.17-21 In the combustion model, laminar flame speed is calculated by using the equations developed by Gu¨lder.22 The geometrical features of the flame front are computed from a geometric submodel installed on the basics17,18 by using the mathematical relations given in the literature.23 Further details on the combustion model and its submodels can be found in Sezer.24 2.2. Computation of Cycle. As known, SI engine cycles consist of four consecutive processes: intake, compression, expansion, and exhaust. Intake and exhaust processes are computed by using the approximation method given by Bayraktar and Durgun.21 Compression, combustion, and expansion processes have been computed from the governing equations,25 with arrangment of them for each process in a suitable manner. Further details of the cycle model can be found in the literature.24,25 The engine performance parameters, i.e., brake mean effective pressure (bmep) and brakespecific fuel consumption (bsfc), have also been determined from the well-known equations.21,24,25 2.3. Exergy Analysis. The second law of thermodynamics has been applied to the above model for the exergetic computations. As known, the second law can be stated by the statement of entropy balance26

∆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 laws of 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 temperature of the dead state. Availability is defined as the maximum theoretical work that can be obtained from a combined system (combination of a system and its 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 (17) Blizard, N. C.; Keck, J. C. Experimental and theoretical investigation of turbulent burning model for internal combustion engines. SAE Paper 740191, 1974. (18) Keck, J. C. Turbulent flame structure and speed in spark-ignition engines. International 19th Symposium on Combustion, 1982; pp 14511466. (19) Tabaczynski, R. J.; Ferguson, C. R.; Radhakrishnan, K. A turbulent entrainment model for spark-ignition combustion. SAE Paper 770647, 1977. (20) Tabaczynski, R. J.; Trinker, F. H.; Sahnnon, B. A. S. Further refinement 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-ignition engine cycles. Energy Sources 2003, 25, 651–666. ¨ . Correlations of laminar combustion data for alternative (22) Gu¨lder, O SI engine fuels. SAE Paper 841000, 1984. (23) Bilgin, A. Geometric features of the flame propagation process for an SI engine having dual-ignition system. Int. J. Energy Res. 2002, 26, 987–1000. (24) Sezer, I. Application of exergy analysis to spark ignition engine cycle. Ph.D. Thesis, Karadeniz Technical University, Trabzon, Turkey, 2008. (25) Ferguson, C. R. Internal Combustion Engine Applied Thermosciences; John Wiley and Sons: New York, 1985. (26) Moran, J. M.; Shapiro, H. N. Fundamentals of Engineering Thermodynamics, 3rd ed.; John Wiley and Sons: New York, 1998. (27) Rakopoulos, C. D.; Giakoumis, E. G. Second-law analyses applied to internal combustion engines operation. Prog. Energy Combust. Sci. 2006, 32, 2–47.

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of two contributions: thermomechanical exergy, Atm, and chemical exergy, Ach.28 Thermomechanical exergy is defined as the maximum extractable work from the combined system, as the system comes into thermal and mechanical equilibrium with the environment, and it is given as in eq 3

Atm ) E + p0V - T0S -

∑µ

0,imi

∑ m (µ i

I˙comb ) T0S˙comb

(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 thermal and mechanical equilibrium with the environment and no work potential exist between the system and environment because of temperature and pressure differences. However, the system does not reach the chemical equilibrium with the environment, because the contents of the system are not permitted to mix with the environment or enter the chemical reaction with environmental components. In principle, the difference between the compositions of the system at the restricted dead state conditions and the environment can be used to obtain additional work.28 The maximum work obtained in this way is called chemical exergy, which can be determined as

Ach )

The quantities of h′, c′, o′, s′, and w′ in eqs 7 and 8 represent mass fractions of elements carbon, hydrogen, oxygen, sulfur, and the water content in the fuel, respectively. The last term on the right-hand side of eq 6 illustrates exergy destruction because of combustion. It is calculated as24,29

0 0,i - µi )

where S˙comb is the rate of entropy generation because of combustion irreversibilities. It is calculated for two-zone combustion model as

S˙comb )

∆A ) A2 - A1 ) AQ - AW - Adest

(5)

where ∆A is the change of the total system availability for a given process, A2 is the total availability at the end of the process, A1 is the total availability at the start of the process, AQ is the availability transferred accompanying the heat transfer, AW is the availability transfer because of work, and Adest is destroyed availability by irreversible processes. Considering fuel chemical exergy, the balance equation for the engine cylinder is written as24,29

) (

T0 dQ mf dxb dIcomb dAtot dW dV ) 1- p0 a + dθ T dθ dθ dθ mtot dθ f,ch dθ (6)

)

The left-hand side of eq 6 is the rate of change in total exergy of cylinder contents. The first and second terms on the right-hand side represent exergy transfer with heat and exergy transfer with work, respectively. The third term on the right-hand side corresponds to the burned fuel exergy, where mf and mtot are the masses of 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.

[

o′ h′ + 0.0432 + c′ c′ s′ h′ 0.2196 1 - 2.0628 c′ c′

Af,ch ) af,chmf ) QLHV 1.0401 + 0.01728

(

)]

(10)

where mb and mu are burned and unburned masses of the cylinder contents and sb and su are entropy values of the burned and unburned gases, respectively. Moreover, total exergy destruction considered here consists of combustion and heat transfer irreversibilities as follows:24,29

I˙tot ) I˙comb + I˙Q

(11)

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

S˙Q )

µ0i

(

d(mbsb) d(musu) + dθ dθ

(4)

where 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

(9)

˙b Q ˙u Q + Tb Tu

(12)

˙ b and Q ˙ u are the rates of heat loss from the burned and where Q unburned zones at temperatures Tb and Tu, respectively. The efficiency is defined to compare different engine size applications or evaluate effects of various improvements from the perspective of the first- or second-law analysis.27 The first-law (or energy-based) efficiency is defined as24,25

ηI )

W energy out (as work) ) energy in mfQLHV

where W is indicated work output. Various second-law efficiencies (exergetic efficiency or effectiveness) have been defined in the literature.26,27 The equation given below is used in this study for the second-law efficiency

ηII )

AW exergy out (as work) ) exergy in mfaf,ch

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

QLHV ) [33.91c′ + 125.6h′ - 10.89(o′ - s′) - 2.51(9h′ - w′)] (8) (28) Alkidas, A. C. The application of availability and energy balances to a diesel engine. J. Eng. Gas Turbines Power 1988, 110, 462–469. (29) Zhang, S. The second law analysis of a spark ignition engine fueled with compressed natural gas. M.S. Thesis, University of Windsor, Ontario, Canada, 2002. (30) Kotas, T. J. The Exergy Method of Thermal Plant Analysis; Krieger Publishing: Malabar, FL, 1995.

(14)

where AW is exergy transfer with work.

3. Numerical Applications 3.1. Computer Program and Solution Procedure. A computer program has been written to perform the numerical computations concerning the presented model. Exergetic calculations were performed simultaneously depending upon the thermodynamic state of the cylinder content. The results obtained have been corrected using the following equations:25 ε1 ) 1 - (Vm/V)

(7)

(13)

(44)

ε2 ) 1 + [W/∆(mu) + Qw]

(45) 10-4

Selecting the values of ε1 and ε2 at the levels of was found confidently, and the computer program was terminated. 3.2. Validation of the Model. To demonstrate the validation of the present cycle model, the predicted values from the model are compared to experimental data obtained from the literature.20,31 These comparisons are given in parts a and b of Figure 1 for the conditions specified on the figure, and engine specifications are given in Table 2. The burned mass fraction and cylinder (31) Rakopoulos, C. D. Evaluation of a spark ignition engine cycle using first and second law analysis techniques. Energy ConVers. Manage. 1993, 34 (12), 1299–1314.


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Figure 1. Comparison of predicted values with experimental data. Table 2. Specifications of the Enginesa specification

r

D Xs (mm)

S (mm)

Lcr div Liv (mm) (mm) (mm)

9.9 0.0 83 74 122 experimental engine I20 experimental engine II31 7 0.1 76.2 111.125 220 parametric study engine25 10 0.0 100 80 160

33 30 40

4.4 4.2 5

a The parameters not present in the references, i.e., d and L , are iv iv determined from the mathematical relations given by Heywood.32

pressure are selected as comparison parameters. As seen in the figures, predictions are in good agreement with the experimental data. Therefore, it can be said that the presented model has an enough level of confidence for analysis of engine performance and parametric investigation. 4. Results and Discussion After the present model has been constructed and controlled, the exergetic computations were performed. The engine specifications25 for the parametric investigation devoted to exergy analysis are given in Table 2. Parts a-f of Figure 2 show the variations of exergetic terms with respect to crank angle for examined fuels. When using methanol and ethanol, a larger amount of fuel is required for supplying the same amount of energy in the cylinder because (32) Heywood, J. B. Internal Combustion Engine Fundamentals; McGraw-Hill: New York, 1988.

of their lower calorific values and also the lower stoichiometric air/fuel ratios as seen in Table 1. Therefore, equivalence ratios are organized to supply the same exergy entry into the cylinder. Thus, equivalence ratios are equal to 0.92 for methanol and 0.95 for ethanol, while it is 1.0 for isooctane as seen Table 3. Exergy can be transferred to or from a system accompanying heat and work, similar to energy, but it can be also destroyed by irreversibilities during a process differently from energy.26 The variations of exergy transfer with heat are shown in Figure 2a for three fuels. The negative values in this figure show that the direction of exergy transfer is from cylinder contents to the cylinder walls, which means that some amount of the exergy is lost. Isooctane has the highest values in magnitude, while methanol gives the lowest one. These variations can be attributed to both combustion temperatures and combustion durations of the fuels. As seen in Table 3, isooctane has a higher combustion temperature because of a higher heating value than those of methanol and ethanol. Further, isooctane has a longer combustion duration compared to methanol and ethanol, which extends the contact duration of hot gases with cylinder walls. Both of these effects cause an increase in the amount of heat transferred and, thus, the exergy transfer with heat, as expected. Variations in exergy transfer with work are given in Figure 2b. In this figure, the negative part of the variations shows that the direction of work transfer is from the piston to the cylinder contents. Coversely, the positive part of the variations shows that work transfer is from the cylinder contents to the piston, which corresponds to the useful work. Three fuels examined have almost the same values during compression (before 0 CAD, i.e., before top dead center). However, isooctane gives the maximum values, and ethanol is better than methanol during expansion. These variations in exergy transfer with work can be explained depending upon the calorific values of fuels. As mentioned in the Introduction and seen in Table 1, the lower calorific values of methanol and ethanol cause a decrease in work output and exergy transfer with work, naturally, for methanol and ethanol. Variations in exergy transfer with work are also suitable for the bmep values given in Table 3. Figure 2c shows the variations of irreversibilities, which have reached the minimum values for methanol. Ethanol gives higher values than methanol, and isooctane has the greatest irreversibilities as seen in the figure. The reductions in irreversibilities for oxygenated fuels can be based on the simple molecular structure of these fuels as explained in the literature.16 Additionally, the shorter combustion durations for oxygenated fuels, as seen in Table 3, provide another contribution to the reduction of irreversibilities as cited in the literature.33 It is also noted that both of these positive effects on irreversibilities are dominant over the negative effect of lower combustion temperatures obtained with oxygenated fuels. Figure 2d shows the thermomechanical exergy variations for the examined fuels. The variations in thermomechanical exergy reflect the combination of exergy transfers accompanying heat and work and also contain irreversibilities, as shown in eq 6. Isooctane has the lower values than methanol and ethanol during combustion. The faster burning of oxygenated fuels as cited in the Introduction causes the rising of temperature and pressure in the cylinder earlier for methanol and ethanol in comparison to isooctane. For this reason, thermomechanical exergy values are a bit higher for methanol and ethanol than that of isooctane during combustion. However, methanol and ethanol give very (33) Caton, J. A. On the destruction of availability (exergy) due to combustion processes with specific application to internal combustion engines. Energy 2000, 25, 1097–1117.


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

close but the lower thermomechanical exergy values than those of isooctane during expansion. These variations in thermomechanical exergy values during expansion can be based on the previous statements of exergy transfers with heat and work and also irreversibilities. The completion of combustion for oxygenated fuels in advance results in lower thermomechanical values for methanol and ethanol toward the end of expansion in comparison to isooctane.

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


Sezer et al.

Table 3. Values Obtained from the Cycle Model for Examined Fuels fuels

φ

mf (g)

mtot

Qf (J)

Af,ch

Tmax bmep ∆θb (K) (bar) (deg)

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

103 78 69

fuel exergies for the three fuels examined at the end of the combustion period because of the lean mixtures for oxygenated fuels and stoichiometric mixture for isooctane. Variations in total exergy values were given in Figure 2f. Total exergy is a summation of fuel exergy and thermomechanical exergy, as shown in eq 6. Therefore, it reflects the characteristics of both of them. For example, the fuel exergy is constant before the start of combustion, and, the increase in total exergy comes from the increase in thermomechanical exergy. At the end of the expansion stroke, fuel exergy decreases to zero and the exergy transfer with exhaust comes from the thermomechanical exergy. According to this variation, ethanol gives the lowest total exergy values, while isooctane has the highest ones during expansion. Exergy transfer with heat reaches the greatest value for isooctane at the end of the expansion as seen in the figure. Methanol also gives very close but higher values than ethanol; thus, ethanol has the least exergy transfer associated with exhaust. Figure 3a shows the distributions of fuel exergy on various exergetic terms for the different fuels examined. The percentages of irreversibilities are lower for oxygenated fuels in comparison to isooctane because of explanations above. Methanol gives the least amount of irreversibilities among examined fuels. The distributions of irreversibilities are about 17.53, 18.16, and 18.97% for methanol, ethanol, and isooctane, respectively. The percentages of exergy transfer with heat are very close for all examined fuels, but oxygenated fuels have lower values than isooctane. The distributions of exergy transfer with heat are about 7.71, 7.97, and 8.52% for methanol, ethanol, and isooctane, respectively. The proportion of exergy transfer with work is the maximum for isooctane, while the minimum value is obtained with methanol, as suitable to exergy variations given above. The distributions of exergy transfer with work are about 35.61, 37.19, and 38.44% for methanol, ethanol, and isooctane, respectively. The percentages of exergy transfer with exhaust are smaller for methanol and ethanol in comparison to isooctane as seen in the figure. The distributions of exergy transfer associated with exhaust gases are about 34.15, 33.21, and 34.25% for methanol, ethanol, and isooctane, respectively. Variations of the first-law efficiency (ηI), second-law efficiency (ηII), and brake-specific fuel consumption (bsfc) values were given in Figure 3b. The first- and second-law efficiencies are similar in characteristics but different in magnitude as seen in the figure. The first-law efficiency is the maximum for isooctane because of maximum work output. Ethanol has a lower efficiency value than isooctane, and methanol gives the minimum first-law efficiency. Similarly, isooctane has the maximum second-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 interesting variations in terms of bsfc when considering the variations in efficiencies. Isooctane has given the least amount of bsfc, while methanol has the greatest one. The increments in bsfc are more than twice for methanol and more than one and half times for ethanol when compared to isooctane. Additionally, it is noted here that methanol and ethanol have higher research octane numbers than isooctane as seen in Table 1, which allows for

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

selection of the higher compression ratios for them. Hence, it is possible to improve engine performance for oxygenates by selecting higher compression ratios as suggested in the literature.8,15 5. Conclusions In this study, the use of oxygenated fuels in SI engines has been evaluated by using a two-zone quasi-dimensional cycle model via exergy analysis. The following conclusions can be drawn from the present study: (1) The quasi-dimensional model of the engine cycle is very suitable for parametric studies, and coupling of this model with exergy analysis is very useful in assessment of engine operation. (2) Methanol and ethanol are suitable from the perspective of exergy analysis because of the less 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 to isooctane. Exergy transfer with heat decreases about 9.47% for methanol and 6.45% for ethanol in comparison to isooctane. Additionally, exergy transfer with exhaust decreases in magnitude about 0.37 and 3.07% for ethanol and methanol, respectively, in comparison to isooctane. (3) The oxygenated fuels examined result in a decrease in the work output because of 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 efficiencies than isooctane. The decrements are about 0.43 and 0.05%, respectively, for methanol and ethanol. The second-law efficiencies also decrease about 2.83% for methanol and 1.23%


for ethanol in comparison to isooctane. However, it is noted that engine efficiency and performance can be increased by selecting higher compression ratios because of the higher octane quality of oxygenates. (5) The brake-specific fuel consumption is quite higher for especially methanol and ethanol in comparison to isooctane. The increments in bsfc are about 132.2% for methanol and 65.5% for ethanol in comparison to isooctane. Considering these extremely higher bsfc values for oxygenated fuels, the use of methanol and ethanol as fuel additives in gasoline seems more reasonable. Thus, it is possible to decrease the irreversibilities by improving the combustion properties of gasoline without increased bsfc. Nomenclature A ) availability or exergy, J Aexh ) exergy transfer associated with exhaust gases, J Af,ch ) fuel chemical exergy, J AQ ) exergy transfer with heat, J Atm ) thermomechanical exergy, J AW ) exergy transfer with work, J Atot ) total exergy, J bmep ) brake mean effective pressure, bar bsfc ) brake-specific fuel consumption, g kW-1 h-1 div ) intake valve diameter, m D ) cylinder diameter, m E ) energy, J f ) mass residual gas fraction, dimensionless I ) irreversibilities, J Lcr ) connecting rod length, m Liv ) maximum intake valve lift, m m ) mass, kg p ) pressure, bar R ) cylinder radius, m Rs ) radius of spark plug location from cylinder axis, m QLVH ) lower heating value (or calorific value) of fuel, kJ/kg Qw ) total heat transfer to cylinder walls, J

Energy & Fuels, Vol. 23, 2009 1807 T ) temperature, K r ) compression ratio, dimensionless S ) entropy, J U ) internal energy, J xb ) mass fraction burned, dimensionless Xs ) Rs/R ) spark plug location, dimensionless AbbreViations CAD ) crank angle degree Greek Letters φ ) fuel/air equivalence ratio, dimensionless ηI ) the first-law efficiency, % ηII ) the second-law efficiency, % θ ) crank angle, CAD θs ) spark timing crank angle, CAD ∆θb ) burn duration, CAD Subscripts 0 ) reference or dead state conditions b ) burned ch ) chemical cr ) connecting rod comb ) combustion dest ) destruction exh ) exhaust f ) fuel kin ) kinetic max ) maximum pot ) potential s ) spark tm ) thermomechanical tot ) total u ) unburned w ) wall EF8002608

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

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