Energy and exergy analyses of a gasoline engine

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1 INTERNATIONAL JOURNAL OF ENERGY RESEARCH Published online in Wiley InterScience ( Energy and exergy analyses of a gasoline engine C. Sayin 1, M. Hosoz 2, *,y, M. Canakci 2 and I. Kilicaslan 2 1 Department of Mechanical Education, Marmara University, Istanbul, 34722, Turkey 2 Department of Mechanical Education, Kocaeli University, Kocaeli, 41380, Turkey SUMMARY This study presents comparative energy and exergy analyses of a four-cylinder, four-stroke spark-ignition engine using gasoline fuels of three different research octane numbers (RONs), namely 91, 93 and Each fuel test was performed by varying the engine speed between 10 and 2400 rpm while keeping the engine torque at and 40 Nm. Then, using the steady-state data along with energy and exergy rate balance equations, various performance parameters of the engine were evaluated for each fuel case. It was found that the gasoline of 91-RON, the design octane rating of the test engine, yielded better energetic and exergetic performance, while the exergetic performance parameters were slightly lower than the corresponding energetic ones. Furthermore, this study revealed that the combustion was the most important contributor to the system inefficiency, and almost all performance parameters increased with increasing engine speed. KEY WORDS: gasoline engine; octane number; energy analysis; exergy analysis 1. INTRODUCTION The performance, fuel consumption and exhaust emissions of internal combustion (IC) engines have been improved considerably since the introduction of these engines in the 19th century. These improvements have been brought about by not only advancing the engine design but also employing fuels with better properties. Because combustion is strongly affected by the fuel properties, both the engine performance and exhaust emissions vary as functions of them. On the other hand, the quality of gasoline, the main fuel for gasoline engines, is usually indicated by the research octane number (RON). The RON of a gasoline is a measure of its resistance to detonation. The engine design and compression ratio are the main factors determining the RON *Correspondence to: M. Hosoz, Department of Mechanical Education, Kocaeli University, Umuttepe-Izmit, 41380, Kocaeli, Turkey. y mhosoz@kou.edu.tr Contract/grant sponsor: TOFAS Automotive Plant and Turkish Petroleum Refinery Corporation (TUPRAS) Received 13 April 06 Revised 22 June 06 Accepted 28 June 06

2 C. SAYIN ET AL. of gasoline required by an engine. Besides them, the weather and driving conditions as well as mechanical conditions of the engine can influence the required RON. The effects of RON on the engine performance and/or exhaust emissions have been recently investigated by many researchers (Batmaz et al., 1997; Sudsanguan and Chanchaowna, 1999; Sakaguchi, 00). On the other hand, the effect of RON on detonation has also been studied (Korkmaz, 1996; Bayraktar, 1997; Twu and Coon, 1998; Mogi et al., 1998). Abdulghani (02) found that as the octane number of the fuel increases the CO and HC emissions decrease but the NO x emission increases. It can be seen from the literature survey that there are only a few studies investigating the effect of higher octane rating on the gasoline engines with carburettor fuel system. Therefore, in the current study, the effect of higher-octane gasoline on the performance and exhaust emissions was studied using the carburettor gasoline engine. Furthermore, the studies in the open literature usually determined various performance parameters of the engine from the perspective of the first law of thermodynamics. However, the first law is inadequate for evaluating some features of energy resource utilization (Moran, 1989; Moran and Shapiro, 00). The thermodynamic details of the operation of a thermal system can be understood better by performing not only an energy analysis but also an exergy analysis of the system. The exergy analysis reveals the locations, causes and magnitudes of energy resource waste in the system. The exergy analysis, also known as second law or availability analysis, of IC engines has been studied by many investigators. Patterson and Van Wylen (1963) studied a thermodynamic cycle simulation which determined availability for the compression and expansion strokes in a spark ignition engine. Flynn et al. (1984) evaluated low-heat-rejection engine concepts in a diesel engine using second law analysis. Primus and Flynn (1986) showed the benefits of using the second law to determine various energy losses in a diesel engine. Van Gerpen and Shapiro (1990) investigated the second law analysis of diesel engine combustion. Alkidas (1998) performed energy and exergy analyses of a diesel engine using experimental data. Lipkea and DeJoode (1998) studied comparative energetic and exergetic performances of two direct injection diesel engines. Alasfour (1997) performed an exergy analysis of a spark ignition engine to evaluate the use of a butanol gasoline blend as fuel. Caton (00) reviewed over two dozen previous studies on the exergetic performance of IC engines. Canakci and Hosoz (06) presented a comparative energy and exergy analyses of a four-cylinder turbocharged diesel engine fuelled with various biodiesels and No. 2 diesel fuel. In this study, energy and exergy analyses of a four-cylinder, four-stroke gasoline engine are investigated. The test engine was powered with gasoline fuels of 91, 93 and 95.3-RONs, and operated in a steady state. The tests for each fuel type were conducted by varying the engine speed and torque. Using the experimental data, the reaction equations, the energy rate balance and the exergy rate balance for the engine were determined. Consequently, various energetic and exergetic performance parameters of the engine were evaluated for each RON case and compared with each others. 2. EXPERIMENTAL SET-UP AND TESTING PROCEDURE The tests were performed on a four-cylinder, four-stroke Fiat DKS 1.6 L spark-ignition engine, which originally requires 91-RON gasoline. The schematic layout of the experimental apparatus is shown in Figure 1, and details of the engine specifications are shown in Table I. The engine

3 ENERGY AND EXERGY ANALYSES OF A GASOLINE ENGINE Figure 1. Schematic layout of the experimental apparatus. Table I. The technical specifications of the Fiat DKS engine. Engine type 131 A 1016 Cylinder number 4 Cylinder bore 84 mm Stroke 71.5 mm Firing order Static ignition timing ( o ) 5 BTDC Intake valve opens ( o ) 42 BTDC Total cylinder volume 1600 cm 3 Compression ratio 8:1 Maximum torque Nm at 3400 rpm Maximum power kw Engine octane requirement 91 RON was fuelled with commercial grade gasolines of 91, 93 and 95.3-RONs. Compositions and some important properties of the tested fuels are given in Table II. The engine torque was measured with a hydraulic dynamometer having a range of Nm. The flow rate of the intake air was measured using a conic edge orifice metre coupled to an inclined manometer. The fuel consumption of the engine was determined by monitoring the fuel level in a measurement container. The engine speed was continuously measured using a photoelectric tachometer. The carbon monoxide (CO) and unburned hydrocarbon (HC) emissions were measured using a Bilsa 3 model exhaust gas analyser, which was calibrated before each test. The intake air and exhaust temperatures were measured using type K thermocouples. In each fuel test, the engine speed was varied between 10 and 2400 rpm with intervals of 400 rpm, while the engine torque was kept at and 40 Nm for each engine speed. Data were

4 C. SAYIN ET AL. Table II. Some properties of the fuels used in the tests (TUPRAS, 1996). Research octane number Typical formula C 6.93 H C 6.95 H C 6.97 H Average molecular weight (kg kmol 1 ) Lower heating value (kj kg 1 ) Specific exergy (kj kg 1 ) (calculated from Equation (10)) collected to analyse the performance of the engine after achieving the steady-state operation. Each test run was repeated three times and the results of the three repetitions were averaged. Before experiments, the engine was adjusted to its recommended catalogue values. Further details of the experimental set-up and testing procedure can be found in Sayin et al. (05). 3. DETERMINING THE REACTION EQUATIONS Energy and exergy analyses of the test engine for each test operation can be performed after determining the reaction equation for the considered operation. For this purpose, it is necessary to make some simplifying assumptions. Accordingly, it is assumed that there is no water vapour in the combustion air, and the air contains 21% oxygen and 79% nitrogen on a molar basis. Moreover, it is also assumed that the nitrogen in the air cannot undergo chemical reactions to form NO and NO 2. After these assumptions, the general form of the reaction equation is C x H y þ aðo 2 þ 3:76N 2 Þ!bO 2 þ cco þ dco 2 þ ec x H y þ f N 2 þ gh 2 O ð1þ In all test operations, the fuel chemical formulas, fuel and air flow rates and the flow rates of the exhausted CO and HC are known. Thus, by applying conservation of mass principle to the carbon, hydrogen, oxygen and nitrogen, the unknown coefficients in Equation (1) can easily be determined. The energy and exergy analyses of the engine are performed using these coefficients. 4. ENERGY ANALYSIS With the purpose of simplifying the first law calculations, the following assumptions were made: * The engine runs at a steady state. * The combustion air and exhaust gas each forms ideal gas mixtures. * Potential and kinetic energy effects of the combustion air, fuel stream and exhaust gas are ignored. After these assumptions, the energy rate balance on a per mole of fuel basis for a control volume consisting of the entire engine including the dynamometer can be written as Q cv n F W cv n F ¼ %h P %h R ¼ X P n out ð%h 0 f þ D %hþ out X R n in ð%h 0 f þ D %hþ in ð2þ

5 ENERGY AND EXERGY ANALYSES OF A GASOLINE ENGINE where n F is the molar flow rate of the fuel, and %h P and %h R represent, respectively, the enthalpies of the products and reactants per mole of fuel. On the right side, out and in denote outgoing products and incoming fuel and air streams, while the coefficients n out and n in are the coefficients of the reaction equation, whose values were calculated before. Finally, %h 0 f is enthalpy of formation and D%h represent enthalpy change owing to a change of state at constant composition. For a considered reactant or product at temperature T, the enthalpy change term in Equation (2) can be evaluated from D%h ¼ %hðtþ %hðt ref Þ ð3þ The tables in Moran and Shapiro (00), and Szargut et al. (1998) can be used to get the formation enthalpies and enthalpy changes for the air and combustion products. Then, the formation enthalpies of the tested fuels can be determined from the general reaction equation for the complete combustion with the theoretical amount of air given below: C x H y þ aðo 2 þ 3:76N 2 Þ!bCO 2 þ ch 2 O þ dn 2 ð4þ The values of the unknown coefficients a, b, c and d in Equation (4) can be determined by applying the conservation of mass principle to each element with the use of the fuel chemical formulas given in Table II. Then, the enthalpy of combustion for each fuel, which is also termed as the lower heating value, can be determined from %h 0 RP ¼ LHV ¼ bð %h 0 f Þ CO 2 þ cð%h 0 f Þ H 2 OðgÞ þ dð%h 0 f Þ N 2 ð%h 0 f Þ F að%h 0 f Þ O 2 3:76að%h 0 f Þ N 2 ð5þ where the subscript F stands for fuel and g indicates that the water in the products is a vapour. Formation enthalpies of O 2 and N 2 are equal to zero. Then, taking the lower heating value of the considered fuel from Table II, formation enthalpy of each fuel can be determined from Equation (5). These enthalpies for gasoline fuels of 91, 93 and 95.3-RONs were calculated as , 83 and 1367 kj kmol 1, respectively. The heat flow rate to the control volume can be determined by inserting the values for the fuel molar flow rate, brake power, enthalpy terms and the coefficients of the reaction into Equation (2). The energy input accompanying the combustion air can be ignored since the combustion air is very close to the standard reference state, defined by T ref ¼ 8C and P ref ¼ 1 atm: Thus, the energy loss due to the exhaust gas, i.e. exhaust loss, can be evaluated from Q ex ¼ n F jlhvj W cv j Q cv j ð6þ The brake thermal efficiency (BTE), which is defined as the ratio of the power output to the fuel energy input, indicates the first law performance of the engine, i.e.: BTE ¼ W cv n F jlhvj ð7þ The results of the heat flow rate to the control volume, exhaust loss and BTE calculations are presented in Section 6.

6 C. SAYIN ET AL. 5. EXERGY ANALYSIS Exergy is the maximum theoretical work which can be obtained when a system of interest interacts with a reference environment to equilibrium (Dincer, 00). The order of exergy destructions and losses in the processes and components of a thermal system can be revealed by the exergy analysis of the system. The results of exergy analysis can be used for pinpointing the processes in a thermal system on which further studies must be concentrated for better energy source utilization. In this study, the exergy calculations were performed in relation to the reference environment having a temperature (T 0 ) of K and a pressure (P 0 ) of 1 atm. Moreover, it is supposed that the reference environment consists of an ideal gas mixture with the following composition on a molar basis: N 2, 75.67%; O 2,.%; CO 2, 0.03%; H 2 O, 3.12%; other, 0.83%. The specific flow exergy of a fluid stream can be found by summing thermomechanical and chemical exergies, i.e.: %e ¼ %e th þ %e ch ð8þ The thermomechanical exergy can be defined as %e th ¼ %h %h 0 T 0 ð%s %s 0 Þ ð9þ where %h and %s signify the specific enthalpy and entropy of the fluid, respectively, whereas %h 0 and %s 0 stand for the corresponding values of these properties when the fluid comes to equilibrium with the reference environment. The specific chemical exergies of liquid fuels on a unit mass basis can be determined from Kotas (1995) e ch F ¼ 1:0401 þ 0:1728 h c þ 0:0432 o c þ 0:2169 s c 1 2:0628 h c jlhvj where h, c, o and s are the mass fractions of H, C, O and S, respectively. In this study, the chemical exergies of the fuels were calculated using Equation (10) and Table II. The exhaust gas can be assumed as a mixture of ideal gases. Then, the thermomechanical exergy of the exhaust gas at the temperature T and pressure p, and containing n components can be obtained from %e th ¼ Xn a i %h i ðtþ %h i ðt 0 Þ T 0 %s 0 ðtþ %s 0 ðt 0 Þ %R ln p ð11þ p i¼1 0 where a i is the molar amount of the component i, %s 0 is the absolute entropy at the standard pressure and %R is the universal gas constant. The chemical exergy of the exhaust gas is X %e ch ¼ %RT n 0 a i ln y i ð12þ i¼1 where y i is the molar ratio of the ith component in the exhaust gas and y e i is the molar ratio of the ith component in the reference environment. Both thermomechanical and chemical exergies of the combustion air can be ignored since the intake air was very close to the reference state in the all test operations. Likewise, because its properties were almost equal to those in the reference conditions, the thermomechanical exergy y e i ð10þ

7 ENERGY AND EXERGY ANALYSES OF A GASOLINE ENGINE of the fuel can also be ignored. Considering the coefficients in the reaction equation, the specific flow exergy of the exhaust gas per mole of fuel can be evaluated by summing the results of Equations (11) and (12). The exergy rate balance for the engine operating at a steady state can be expressed as 0 ¼ X ð1 T 0 =T j Þ Q cv W cv þ n F%e F n F%e ex E d ð13þ j where %e F and %e ex are specific exergies of the fuel and exhaust gas, respectively, and T j is the absolute temperature of the boundary section from where the heat is rejected. The term ð1 T 0 =T j Þ Q cv indicates the rate of exergy transfer accompanying heat. The term n F%e F accounts for the rate of exergy entering the control volume accompanying the fuel stream, while n F%e ex describes the exergy leaving the control volume accompanying the exhaust gas. Finally, the term E d signifies the rate of exergy destroyed in the control volume owing to irreversibility. In this study, calculations of the rate of exergy transfer accompanying heat is based on the assumption that Q cv is transferred from the control volume boundary having the same temperature as the engine coolant leaving the engine block. Then, inserting the values for the rate of exergy transfer accompanying heat, the rates of exergy accompanying fluid streams and brake power into Equation (13), the rate of exergy destroyed in the engine can be evaluated. Finally, the exergetic efficiency of the engine is given by the ratio of the power output to the fuel exergy input, i.e.: W e ¼ cv ð14þ n F%e F The results of the exergy calculations are presented in the next section. 6. RESULTS AND DISCUSSION The fuel energies ð n F LHVÞ entering the engine are indicated in Figure 2. For any selected combination of engine speed and torque, the energy provided by the gasoline fuel to obtain the same output power decreases with decreasing octane rating. When a fuel with an octane rating higher than required by the engine is used, this causes a longer ignition delay and a shorter flame. Consequently, this phenomenon reduces the maximum pressure and output power. In other words, a higher amount of fuel energy is required to produce a certain output power. Therefore, the operations with 91-RON gasoline, the design octane rating of the test engine, required the lowest fuel energy. Furthermore, increasing engine speed and torque causes an increase in the fuel consumption, thus raising the fuel energy entering the engine. The BTEs of the engine are shown in Figure 3. Because the power output for a selected combination of engine speed and torque is constant, the BTE for the selected combination is a function of only the fuel energy. In this study, the fuel energy entering the engine increased with increasing octane rating. Therefore, the BTE decreased with increasing octane rating. If the necessary adjustment such as ignition timing was made on the engine in accordance with the increasing octane number, higher efficiencies could be obtained. However, such an adjustment was not made. Furthermore, the BTE first increased with increasing engine speed, and then tended to decrease. Starting at the maximum BTE value, increasing or decreasing speed at constant torque causes a decrease in the engine volumetric efficiency, thereby decreasing the

8 C. SAYIN ET AL. Fuel energy (kw) RON, Nm 93-RON, Nm 95.3-RON, Nm Figure 2. Fuel energies entering the engine as functions of engine speed and torque. BTE (%) RON, Nm 93-RON, Nm 95.3-RON, Nm Figure 3. Brake thermal efficiencies as functions of engine speed and torque. BTE. Volumetric efficiency is a measure of the rate of combustion air. As the volumetric efficiency decreases, fuel flow rate will decrease for a constant F/A ratio, thus resulting in a decrease in the power output of the engine. The main reason of the increase in BTE with the engine speed at low speeds is the decreased amount of fuel supply to the engine. However, increased mechanical frictions dominate this positive effect after a certain speed, and BTE starts to decrease. The heat flow rates from the control volume are shown in Figure 4. The operations using fuels with higher octane ratings result in higher rates of heat flow from the control volume. Because the use of higher octane rating gasoline yields lower CO and unburned HC concentrations in the exhaust gas, enthalpy of combustion products becomes lower, i.e. absolute value of negative h p

9 ENERGY AND EXERGY ANALYSES OF A GASOLINE ENGINE increases. Therefore, a higher rate of total energy in the forms of heat and work is transferred from the control volume in the operations with higher octane ratings. Because the engine power output is the same for a selected combination of engine speed and torque, a higher heat flow rate to the control volume occurs in the operations with higher octane rating. The heat flow increases with increasing engine speed and torque due to the increased fuel energy entering the engine. Although the heat flow rate from the engine can be reduced by insulating the walls of the combustion chamber, this causes an increase in the temperature of the exhaust gas, thus increasing the energy loss due to the exhaust gas (Parlak et al., 05). The energy losses due to the exhaust gas as functions of engine speed and torque are indicated in Figure 5. This loss is simply the difference between the fuel energy input and the sum of heat Heat flow rate (kj/s) RON, Nm 93-RON, Nm 95.3-RON, Nm Figure 4. Heat flow rate from the control volume as functions of engine speed and torque. Exhaust loss (kw) RON, Nm 93-RON, Nm 95.3-RON, Nm Figure 5. Energy losses due to the exhaust gas as functions of engine speed and torque.

10 C. SAYIN ET AL. and work transfers from the control volume. Because work transfer is the same for a selected engine speed and torque combination, the exhaust loss is a function of only the difference between fuel energy input and the heat flow rate from the control volume. Since fuel energy input increases faster than the heat flow rate with increasing octane rating, operations with highoctane fuels cause increased exhaust losses. It can be noted that, in the experimental study the exhaust temperature was measured on the external surface of the exhaust manifold. Therefore, the measured temperature was lower than real exhaust gas temperature. Consequently, the calculated exhaust losses were lower than real values. Therefore, according to Equation (2), heat flow rates presented in Figure 4 also contains some amount of exhaust loss originating from the measurement of exhaust temperature on the external surface of the manifold. The fuel exergy inputs ð n F%e F Þ to the engine for the three fuel cases as functions of engine speed and torque are reported in Figure 6. The curves in this figure are comparable to the fuel energy input curves given in Figure 2 because the fuel specific exergy is linked to the lower heating value to some extent. However, the fuel exergy inputs are % higher than the corresponding fuel energy inputs. The exergetic efficiencies of the test engine as functions of engine speed and torque are shown in Figure 7. The exergetic efficiencies of the engine also follow similar trends with the BTEs given in Figure 3. However, the exergetic efficiencies are % lower than the corresponding BTEs because a higher amount of fuel exergy compared to the fuel energy is supplied to the engine. Since the exergetic efficiency takes into account not only the first but also the second law of thermodynamics, it provides a better measure of the performance for a thermal system. The exergy losses accompanying heat flow rate from the engine as function of engine speed and torque are indicated in Figure 8. Since this loss is proportional to the heat flow rate from the engine, the curves given here are analogous to those given in Figure 4. The higher the temperature of the engine surface from where the heat is rejected, the higher the exergy loss accompanying it is. Alkidas (1998) reported that both the exergy loss accompanying heat loss and the rate of exergy destroyed by the combustion could be decreased by insulating the combustion chamber walls of the engine. Fuel exergy (kw) RON, Nm 93-RON, Nm 95.3-RON, Nm Figure 6. Fuel exergies entering the engine as functions of engine speed and torque.

11 ENERGY AND EXERGY ANALYSES OF A GASOLINE ENGINE Exergetic efficiency (%) RON, Nm 93-RON, Nm 95.3-RON, Nm Figure 7. Exergetic efficiencies as functions of engine speed and torque. 7 7 Exergy loss due to heat flow rate (kw) RON, Nm 93-RON, Nm 95.3-RON, Nm Figure 8. Exergy losses accompanying heat flow rate from the control volume as functions of engine speed and torque. The exergy losses due to the exhaust gas as functions of engine speed and torque are reported in Figure 9. It is seen that the curves in this figure follow similar trends with exhaust energy loss curves in Figure 5, although exhaust exergy loss is slightly lower than energy loss. This exergy loss can be decreased by reducing the exhaust gas temperature and the concentrations of CO and unburned HC in the exhaust gas. The rates of exergy destroyed within the engine as functions of engine speed and torque are indicated in Figure 10. Exergy is not conserved, and the irreversible processes in the engine, such as combustion, heat transfer, mixing, friction, etc., destroy a significant fraction of the fuel

12 C. SAYIN ET AL. Exhaust exergy (kw) RON, Nm RON, Nm RON, Nm Figure 9. Exergy losses due to the exhaust gas as functions of engine speed and torque. Exergy destroyed in the engine (kw) RON, Nm 93-RON, Nm 95.3-RON, Nm Figure 10. The rates of exergy destroyed in the engine as functions of engine speed and torque. exergy. The rate of exergy destroyed in the engine increases with increasing octane rating and engine speed. The exergy destruction due to combustion can be reduced by taking some design precautions to increase the combustion temperature such as preheating the intake air and reducing the amount of excess air. However, these precautions may lead to an increase in the exhaust gas temperature, thereby causing a higher exergy loss accompanying the exhaust gas. 7. CONCLUSIONS The energy and exergy analyses of a four-cylinder, four-stroke gasoline engine using the gasoline fuels of 91, 93 and 95.3-RONs have been performed. The steady-state tests for each fuel type

13 ENERGY AND EXERGY ANALYSES OF A GASOLINE ENGINE were conducted by varying the engine speed and torque. Afterwards, the reaction equation, the energy rate balance and the exergy rate balance for each test operation were determined using experimental data. Finally, various energetic and exergetic performance parameters of the engine were evaluated and compared with each others. Considering the results of energy and exergy analyses, it is possible to draw following conclusions: * The engine operates less energy-efficiently when a fuel with an octane rating higher the design rating is used. * Similarly, operations with high octane ratings lead to lower exergetic efficiency, usually a higher exhaust exergy loss and a higher exergy loss due to heat loss. * Because high engine speeds cause a more homogenous mixture and a boosted turbulence in the combustion chamber, almost all energetic and exergetic performance parameters increase with increasing engine speed. * The system inefficiency is mainly caused by the exergy destruction due to the irreversible processes such as combustion. The exergy losses due to the exhaust gas and heat flow from the control volume are other contributors to inefficiency. a i NOMENCLATURE =molar amount of component i BTE =brake thermal efficiency %e =specific molar exergy (kj kmol 1 ) E d =rate of exergy destruction (kw) %h =specific molar enthalpy (kj kmol 1 ) %h 0 f =enthalpy of formation (kj kmol 1 ) %h 0 RP =enthalpy of combustion (kj kmol 1 ) %h 0 =specific molar enthalpy at reference environment (kj kmol 1 ) jlhvj =lower heating value (kj kmol 1 ) n =coefficients in the reaction equation n F =molar flow rate of the fuel (kmol s 1 ) P =pressure (kpa) p 0 =environmental pressure (kpa) Q cv =heat flow rate to the control volume (kj s 1 ) Q ex =energy loss due to the exhaust gas (kw) %R =universal gas constant (kj kmol 1 K 1 ) %s =specific molar entropy (kj kmol 1 K 1 ) %s 0 =specific molar entropy at reference environment (kj kmol 1 K 1 ) T =temperature (K) T 0 =environmental temperature (K) W cv =power obtained from the control volume (kw) y i =molar ratio of the ith component =molar ratio of the ith component in the reference environment y e i Greek symbols e =exergetic efficiency

14 C. SAYIN ET AL. Subscripts ex in out P R ref =exhaust =incoming reactants =outgoing products =product =reactant =reference Superscripts ch th =chemical =thermal ACKNOWLEDGEMENTS The authors would like to thank TOFAS Automotive Plant and Turkish Petroleum Refinery Corporation (TUPRAS) for supporting this project. REFERENCES Abdulghani AA. 02. Effects of octane number on exhaust emissions of a spark ignition engine. International Journal of Energy Research 26(4): DOI: /er.783. Alasfour FN Butanol}a single cylinder engine study: availability analysis. Applied Thermal Engineering 17(6): Alkidas AC The application of availability and energy balances to a diesel engine. Journal of Engineering for Gas Turbines and Power 110(3): Batmaz I, Balci M, Salman S, Erdiller B Experimental analysis of fuel economy and exhaust emissions at petrol engine vehicles. First Automotive Technology Congress, Adana, Turkey, May. Bayraktar H Theoretical investigation of the effect of gasoline-ethanol blends on spark-ignition engine combustion and cycles. Ph.D. Thesis, Karadeniz Technical University, Trabzon, Turkey. Canakci M, Hosoz M. 06. Energy and exergy analyses of a diesel engine fuelled with various biodiesels. Energy Sources, Part B, in press. DOI: / Caton JA. 00. A review of investigations using the second law of thermodynamics to study internal-combustion engines. SAE Paper, No Dincer I. 00. Thermodynamics, exergy and environmental impact. Energy Sources 22(8): DOI: / Flynn PF, Hoag KL, Kamel MM, Primus RJ A new perspective on diesel engine evaluation based on second law analysis. SAE Paper, No Korkmaz I A study on the performance and emission characteristics of gasoline and methanol fuelled sparkignition engines. Ph.D. Thesis, Istanbul Technical University, Istanbul, Turkey. Kotas TJ The Exergy Method of Thermal Plant Analysis. Krieger Publishing Company: Malabar, FL. Lipkea WH, DeJoode AD A comparison of the performance of two direct injection diesel engines from a second law perspective. SAE Paper, No Mogi K, Katsushi H, Arisawa K, Kobayashi H Analysis and avoidance of pre-ignition in S.I. gasoline engines. JSAE Review 19(1):9 14. Moran MJ Availability Analysis. ASME Press: New York. Moran MJ, Shapiro HN. 00. Fundamentals of Engineering Thermodynamics. Wiley: New York. Parlak A, Yasar H, Eldogan O. 05. The effect of thermal barrier coating on a turbo-charged diesel engine performance and exergy potential of the exhaust gas. Energy Conversion and Management 46(3): DOI: / j.enconman Patterson D, Van Wylen G A digital computer simulation for spark-ignited engine cycles. SAE Paper, No Primus RJ, Flynn PF The assessment of losses in diesel engines using second law analysis. In Computer-Aided Engineering of Energy Systems, Vol. 3, Gupta G (ed.), Advanced Energy Systems. ASME: New York,

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