Experimental and Theoretical Investigation of the Effects of Gasoline

Energy & Fuels 2008, 22, 3201–3212

3201

Experimental and Theoretical Investigation of the Effects of Gasoline Blends on Single-Cylinder Diesel Engine Performance and Exhaust Emissions Z. S¸ahin* Karadeniz Technical UniVersity (KTU), Faculty of Marine Science, NaVal Architecture and Marine Engineering, Trabzon 61530, Turkey ReceiVed April 4, 2008. ReVised Manuscript ReceiVed July 9, 2008

The present study investigates experimentally and theoretically the effects of using of gasoline-diesel fuel blends on single-cylinder diesel engine performance and exhaust emissions. The study consists of two parts: (i) Experimental study: The effects of using of different gasoline-diesel fuel blends have been investigated experimentally and compared in a single-cylinder diesel engine. In this part, the effects of 2-10% gasoline blends have been investigated experimentally in a single-cylinder diesel engine at the speeds of 900-1600 rpm and at the selected compression ratios of 18-23. (ii) Theoretical study: A computer program has been developed for the prediction of diesel engine cycles and engine characteristics for the cases of neat diesel fuel (NDF) and gasoline-diesel fuel blends. For calculation of the diesel engine cycle, a quasi-dimensional phenomenological combustion model developed by Shahed and then improved by Ottikkutti has been used and modified with new assumptions. After the engine cycle model for NDF and gasoline blends was proven to give correct results by comparing it to experimental and theoretical results, 2-10% gasoline blends have been investigated theoretically for the same experimental engine. In the theoretical studies, ignition delay, combustion duration, pressure and temperature of the cylinder charge, and mole fraction of carbon monoxide (CO) and nitric oxide (NO) concentrations have been calculated. From experimental results, it is determined that brake-effective power output decreases at the levels of 1.5-4% at low engine compression ratios, such as 18, 19, and 20. However, brake-effective power increases 1.5-4% at high compression ratios, such as 21, 22, and 23. Brake-effective efficiency increases at the levels of 2-6%, and brake-specific fuel consumption decreases at the levels of 2-6% at low ratios of gasoline/diesel fuel. It can be said that the ratios of 4-6% are the most favorable percentage interval of gasoline at the selected compression ratios for this engine. Because the cost of gasoline is higher than that of diesel fuel and the decrease in the brake-specific fuel consumption is low, gasoline blends are not economical for this engine at low compression ratios. However, approximately 3% price saving is attained at low blend percentages for high compression ratios. From theoretical results, it is determined that ignition delay increases as the gasoline ratio increases because of the decreasing cetane number (CN) of the mixture. Combustion duration remains approximately constant with an increasing gasoline ratio at low engine speeds, but it decreases at high speeds. The mole fraction of CO and NO concentrations increase with an increasing gasoline/diesel ratio because of increasing combustion temperatures. As the gasoline/diesel ratio increases, the NO concentration and mole fraction of CO of this engine increase at the levels of 7-29 and 15-90%, respectively.

1. Introduction Direct-injection (DI) diesel engines are widely used as fuelefficient power sources for automotive applications because of their superior fuel economy relative to that of spark-ignition engines and indirect-injection diesel engines at equivalent capacity. However, the increase of diesel fuel prices, stringent emission regulations, and foreseeable future depletion of petroleum reserves force us to research and apply new technologies such as using common-rail systems, fuel injection control strategies, exhaust gas recirculation, fuel-related techniques, etc. to meet the demands of humans for environment and energy. Alternative fuel studies, especially renewable fuel studies, are very important research areas among fuel-related researches. Nowadays, considerable attention has been paid to the develop* To whom correspondence should be addressed. Telephone: 0462-7522805/110. Fax: 0462-752-2158. E-mail: [email protected].

ment of alternative fuel sources in various countries, with particular emphasis on biofuels.1,2 It is well-known that gasoline is not an alternative fuel, and it constitutes the main fuel of spark-ignition engines. However, gasoline gives good results with respect to engine performance characteristics and exhaust emissions when it is used as an additive to diesel fuel using different techniques, such as fumigation and blending. It is important to use fossil fuel sources economically in addition to developing new renewable energy sources. On the other hand, ethanol can be used as another additive, but its production by actual techniques is generally (1) He, B.-Q.; Wang, J.-X.; Yan, X.-G.; Tian, X.; Chen, H. Study on combustion and emission characteristics of diesel engines using ethanol blended diesel fuels. Society of Automotive Engineers (SAE) Tech. Pap. 2003-01-0762, 2003. (2) Rakopoulos, C. D.; Antonopoulos, K. A.; Rakopoulos, D. C.; Hountalas, D. T. Multi-zone modeling of combustion and emissions formation in DI diesel engine operating on ethanol-diesel fuel blends. Energy ConVers. Manage. 2008, 49 (4), 625–643.

10.1021/ef800236y CCC: $40.75  2008 American Chemical Society Published on Web 08/16/2008

3202 Energy & Fuels, Vol. 22, No. 5, 2008

expensive. However, ethanol may be cheaper in some countries, such as Brazil. Moreover, food resources for human beings are very important. For these reasons, gasoline was investigated as blended fuel with diesel fuel in the present study. In the past few decades, researchers have also investigated many methods of using gasoline and alcohols in compression ignition engines to extend diesel fuel supplies and gain some benefits, such as reduced smoke and particulate matter emissions. These techniques can be divided into the following three main categories:1-3 (a) gasoline-diesel fuel blends and alcohol-diesel fuel blends, mixing of the fuels in the fuel tank, displacing up to 25% of the diesel fuel demand; (b) dual injection, separate injection system for each fuel, displacing up to 90% of the diesel fuel demand; and (c) gasoline fumigation or alcohol fumigation, the addition of alcohols or gasoline to the intake air charge, displacing up to 50% of the diesel fuel demand. Obviously, using gasoline-diesel fuel blends or alcohol-diesel fuel blends are the simplest of the abovementioned techniques and do not require any technical modification on the engine side. For these reasons, many relevant experimental studies on alcohol-diesel fuel blends, especially ethanol-diesel fuel blends, have been performed in the literature.1,3-7 He et al.1 have investigated the effect of using of ethanolblended diesel fuels on brake-specific fuel consumption, brakespecific energy consumption, smoke, and NOx emissions in a DI diesel engine. Their results indicate that, with the increase of the ethanol ratio in the blends, smoke reduces significantly, brake-specific energy consumption improves slightly, and combustion duration decreases. On the other hand, the rate of heat release increases but ignition delay, brake-specific fuel consumption, NOx, and unburned ethanol emissions increase. Li et al.4 have investigated experimentally the effects of different ethanol-diesel fuel blends on the performance and emissions of a water-cooled single-cylinder DI diesel engine. In this study, it has been determined that the brake-specific fuel consumption and brake thermal efficiency increase, smoke emission decreases, and CO and NOx decrease with increase of the ethanol content in the blended fuel at the same operating conditions. However, total hydrocarbon emissions increase significantly with an increasing ethanol content. Bilgin et al.5 have studied experimentally the performance of a variable compression ignition engine operated with ethanol-diesel fuel blends. The engine has been operated with ethanol-diesel fuel blends having 2, 4, and 6% ethanol on a volume basis, as well as neat diesel fuel (NDF). Their experimental results indicate that the addition of 4% ethanol to diesel fuel increases brake-effective power output and brake-effective efficiency of the engine, while brake-specific fuel consumption for various compression ratios decreases. It can be seen from the above brief literature survey that there are many experimental studies on the ethanol blends.1,3-7 (3) Abu-Qudais, M.; Haddad, O.; Qudaisat, M. The effect of alcohol fumigation on diesel engine performance and emissions. Energy ConVers. Manage. 2000, 41 (4), 389–399. (4) Li, D.; Zhen, H.; Xingcai, L.; Wu-Gao, Z. Physico-chemical properties of ethanol-diesel blend fuel and its effect on performance and emissions of diesel engines. Renewable Energy 2005, 30 (6), 967–976. (5) Bilgin, A.; Durgun, O.; S¸ahin, Z. The effect of diesel-ethanol blends on diesel engine performance. Energy Sources 2002, 24, 431–440. ¨ .; Çelikten, I.; Usta, N. Effects of ethanol addition on (6) Can, O performance and emissions of turbocharged indirect injection diesel engine running at different injection pressure. Energy ConVers. Manage. 2004, 45 (15-16), 2429–2440. (7) Hansen, A. C.; Zhang, Q.; Lyne, P. W. L. Ethanol-diesel fuel blends. Bioresour. Technol. 2005, 96 (3), 277–285.

S¸ahin

However, theoretical studies on this subject are very scarce.2,8 S¸ahin and Durgun8 have investigated theoretically the effects of using of ethanol-diesel fuel blends in a turbocharged DI diesel engine. They found that, as the equivalence ratio of the ethanol-diesel fuel mixture decreases with an increasing ethanol ratio, the combustion efficiency of the mixture improves. Thus, brake-effective power and brake-effective efficiency increase at the levels of 2 and 4%, respectively, and brake-specific fuel consumption decreases nearly 2%. Because the combustion temperature of the mixture is higher than NDF, the mole fractions of CO and H2 increase. Gasoline fumigation has been studied in a restricted extent in the literature,9-13 but gasoline-diesel fuel blends have not been investigated theoretically or experimentally yet. Kouremenos et al.9 have investigated experimentally gasoline fumigation, and they found that there is a large reduction in smoke with only a slight change in brake-specific fuel consumption at above ∼60% of maximum load. Durgun et al.,10 S¸ahin,11 and S¸ahin et al.12-16 have investigated ethanol and gasoline fumigation both theoretically and experimentally. In these studies, ethanol and gasoline fumigation gave improved results for engine performance and exhaust emissions, and moreover, gasoline fumigation results are more promising than ethanol fumigation results. For example, ethanol and gasoline fumigation have been investigated at the same test engine, and the increase ratios of effective power are at the levels of 4-8 and 4-9% for ethanol and gasoline fumigation, respectively. Detailed information about this subject can be found in previous studies.13,15 Therefore, the present study aims to investigate experimentally and theoretically the effects of using of gasoline-diesel fuel blends in a single-cylinder DI diesel engine on engine performance and exhaust emissions. Investigation of the chemistry effects of the gasoline-diesel fuel blends has not been performed. As a result of this study, technical and (8) S¸ahin, Z.; Durgun, O. Theoretical investigation effects of ethanoldiesel fuel in diesel engine. National Heat Science and Heat Technique Symposium; Karadeniz Technical University: Trabzon, Turkey, 2005; Chapter 15, pp 855-862 (in Turkish). (9) Kouremenos, D. A.; Rakopoulos, C. D.; Kotsiopoulos, P. Comparative performance and emission studies of vaporized diesel fuel and gasoline as supplement in swirl-chamber diesel engines. Energy 1990, 15 (12), 1153– 1160. (10) Durgun, O.; Ayvaz, Y. The use of diesel fuel-gasoline blends in diesel engines. First Trabzon International Energy and EnVironment Symposium; Karadeniz Technical University: Trabzon, Turkey, 1998; Chapter 2, pp 905-912. (11) S¸ahin, Z. Effect of using diesel fuel-light fuel blends on combustion and engine performance in diesel engine. Ph.D. Thesis, KTU Graduate School of Natural and Applied Sciences, Trabzon, Turkey, 2002 (in Turkish). (12) S¸ahin, Z.; Durgun, O. High speed direct injection (DI) light-fuel (gasoline) fumigated vehicle diesel engine. Fuel 2007, 86, 388–399. (13) S¸ahin, Z.; Durgun, O.; Bayram, C. Experimental investigation of gasoline fumigation in a single cylinder direct injection (DI) diesel engine. Energy 2008, 33 (8), 1298–1310. (14) S¸ahin, Z.; Durgun, O. Theoretical investigation of light-fuel fumigation on diesel engine performance and emissions. Energy ConVers. Manage. 2007, 48 (7), 1952–1964. (15) S¸ahin, Z.; Durgun, O. Experimental investigation of ethanol fumigation in a single cylinder diesel engine. The 2nd Energy Conference; University of Ontario: Ottawa, Ontario, Canada, 2006; pp 202-211. (16) S¸ahin, Z.; Durgun, O. Theoretical investigation of light fuel fumigation and probable developments in diesel engines. 14th International Conference on Thermal Engineering and Thermogrammetry (THERMO), Budapest, Hungary, 2005. (17) Durgun, O. Using ethanol-gasoline-isopropanol mixtures in the internal combustion engine. 2nd Combustion Symposium; Istanbul Technical University: Istanbul, Turkey, 1989; pp 325-335 (in Turkish). (18) Holman, J. P. Experimental Methods for Engineers, 7th ed.; McGraw-Hill: New York, 2001.

Single-Cylinder Diesel Engine Performance

Figure 1. Schematic view of the experimental setup. 1, water tank; 2, various fuel tanks; 3, loading resistors; 4, air manometer; 5, engine speed; 6, inlet and outlet temperatures of cooling water; 7, cooling water meter; 8, fuel meter; 9, exhaust temperature; 10, dynamometer; 11, variable compression unit; 12, intake manifold; and 13, engine.

economic aspects of the use of gasoline-diesel fuel blends could be predicted for Turkey as well as other countries. 2. Experimental Study The experiments have been performed at Karadeniz Technical University, Engineering Faculty, Mechanical Engineering Department, Internal Combustion Engines Laboratory. In the present study, a single-cylinder naturally aspirated, four-stoke, water-cooled, variable-compression ratio experimental engine manufactured by Tecquipment is used. The main characteristics of the engine are stroke of 120 mm, bore of 90 mm, displacement of 763.4 cm3, compression ratio interval of 18-24, and injection timing (the start of injection) of 22° before top dead center (TDC). In Figure 1, the schematic view of the experimental setup is shown. The engine is coupled to an electrical dynamometer for loading, and torque output is measured using a spring-weight system. To measure air consumption of the engine, the induction pipe is connected to a large leak-proof air box by means of a flexible hose. At the induction period, air enters into the box through a sharp-edged orifice sited at the box entry. Air consumption is calculated using an inclined manometer and applying the Bernoulli equation. A rotameter flow measuring unit is used to measure the water flow rate, using the calibration curve given by the manufacturing firm. The cooling water inlet and outlet temperatures have been measured using mercury in steel dial thermometers. A thermocouple and pyrometer have been used to measure the engine exhaust temperature. Gasoline-diesel fuel blends have been prepared by mixing them in the fuel tank, and the fuel mixture has been supplied to the engine using the original fuel injection pump. In the present study, the effects of 2-10% gasoline blends have been investigated experimentally at the speeds of 900-1600 rpm and at selected compression ratios of 18-23 for six gasoline ratios. The experiments have been performed at 1:1 gas position (at wide-open throttle). The test procedure can be summarized as follows. The compression ratio has been adjusted to 26 at the beginning of the engine operation. After the engine was warmed, the compression ratio has been set to the desired value. In this study, selected engine compression ratios were 18, 19, 20, 21, 22, and 23. After the compression ratio has been set, the engine speed has been adjusted to 850 rpm, and then it has been raised by 150 rpm increments to 1600 rpm by decreasing the load (engine load can be varied using electronic resistors). First, experiments have been performed for

Energy & Fuels, Vol. 22, No. 5, 2008 3203 NDF, and then, for 2-10% gasoline blends, experiments (that is, 2-10% gasoline by volume in the fuel) have been carried out step by step. The mixtures have been prepared just before the experiments. All experiments have been repeated 6 times for each selected engine compression ratio, engine speed, and gasoline ratio. In the present study, low gasoline ratios are selected and, thus, a small amount of gasoline is used. For this reason, the gasoline-diesel fuel mixture has been supplied to the engine using the original fuel injection pump-injector system. No modification in this fuel system is performed, and fuel injection timing has not been changed during the tests. Ignition advance of this test engine is -22°. All working conditions of the engine for diesel fuel-gasoline blends are taken to be the same as that of neat diesel fuel. Here, only gasoline at the 2-10% volume ratios has been mixed with the diesel fuel in the fuel tank. 2.1. Evaluation of the Experimental Results. In this experimental study, the main measurement values are ambient air temperature and pressure, temperature of exhaust gases, inlet and outlet temperatures of cooling water, wet thermometer temperature of the ambient air, air and cooling water flow rates, engine speed, force of spring dynamometer, and consumption time of ∆V (m3) fuel. Using the properties of the mixture and the above-mentioned measuring values, engine characteristics have been calculated as follows. Here, the main calculation procedure (method) has been summarized. The details of the evaluation of the experimental results and comprehensive information about of the experimental system can be found in Durgun.19,20 Brake-effective power has been calculated as

Ne,1 ) Tbω

(1)

where ω ) πn/30 (s-1) is the angular velocity of the crankshaft, n (rpm) is the rotational speed of the crankshaft, and Tb (N m) is the brake torque. Then, the calculated brake-effective power has been converted to standard atmospheric conditions by taking into accounts the conditions and humidity of the ambient air as follows:19,20

Ne,2 ) 0.1013

Ne,1 T /293xhum p0 √ 0

(2)

where T0 (K) and p0 (MPa) are measured ambient air temperature and pressure, respectively. Humidity correction factor xhum in eq 2 is determined from the dry and wet thermometer temperatures. Brake-specific fuel consumption, be (kg kW-1 h-1), and breakeffective efficiency, ηe, have been calculated using the following well-known relations:

Bt Ne,2

(3)

3600 QLHV,mixbe

(4)

be ) ηe )

where Bt (kg/h) is hourly (total) fuel consumption and QLHV,mix (kJ/ kg) is the lower heating value (LHV) of the fuel mixture. LHVs of each fuel have been determined using the well-known Mendeleyev formula:21

QLHV ) 33.91c′ + 125.6h′ - 10.89(oy′ - s′) - 2.51(9h′ - w′) (5) (19) Durgun, O. Experimental methods and applications diesel engines. Lecture notes, Karadeniz Technical University, Trabzon, Turkey, 2003 (in Turkish). (20) Durgun, O. Experimental methods in the internal combustion engines. Lecture notes, Karadeniz Technical University, Trabzon, Turkey, 1995 (in Turkish). (21) Kolchin, A.; Demidov, V. Design of AutomotiVe Engines; MIR Publishers: Moscow, Russia, 1984.


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Table 1. Properties of Diesel Fuel and Gasoline10-15,22-24 chemical formula molecular mass (kg/kmol) density (kg/m3) lower heating value (kJ/kg) cost (May 2008, Trabzon, Turkey, euro) auto-ignition temperature (°C) vaporization latent heat (kJ/kg) boling point (°C) elemental composition, by mass (%)

diesel fuel

gasoline

C14.342H24.75O0.0495 198.0017 0.810 42 437.34 6.069 254 250 180-360 c′ ) 0.87, h′ ) 0.126, oy′ ) 0.004,

C7H17 101.2130 0.735 45 609.07 6.594 300-400 380-500 25-215 c′ ) 0.831, h′ ) 0.169

In eq 5, c′, h′, oy′, s′, and w′ represent the elemental composition of fuels, and their values for diesel fuel and gasoline have been given in Table 1. In the theoretical computations and the evaluations of experimental results, the chemical formulas of diesel fuel and gasoline are taken approximately as C14.342H24.75O0.0495 and C7H17, respectively.10-16 In the present study, the properties of the gasoline-diesel fuel mixture have been determined using these chemical formulas. Also, in Ferguson’s22 computer code for calculating the spark-ignition engine cycle, the chemical formula of gasoline was taken as C7H17. This is the average closed formula of gasoline, and it does not mean that this fuel consists of only heptane. LHV, density, and stoichiometric air requirement for any mixture can be easily calculated using the following relations:

QLHV,mix )

xdFdQLHV,d + xgFgQLHV,g xdFd + xgFg

(6)

xdFdhmin,d + xgFghmin,g xdFd + xgFg

(7)

hmin,mix )

(8)

2.2. Calculation of Variation Ratios of the Main Engine Characteristics and Cost Analysis. Variation ratios of the main engine characteristics, such as brake-effective power, brake-specific fuel consumption, and brake-effective efficiency, have been calculated in a similar way, for example, as follows:

) ) )

∆Ne Ne,mix - Ne,d 100 (%) ) 100 Ne Ne,d

(9a)

be,mix - be,d ∆be 100 (%) ) 100 be be,d

(9b)

ηe,mix - ηe,d ∆ηe 100 (%) ) 100 ηe ηe,d

(9c)

where Ne,mix and Ne,d are brake-effective powers, be,mix and be,d are brake-specific fuel consumptions, and ηe,mix and ηe,d are brakeeffective efficiencies for the gasoline-diesel fuel blend and NDF, respectively. For the cost analysis, the following relationship developed by Durgun10,17,19,20 has been used:

C2 - C1 ∆C 100 (%) ) 100 (%) C1 C1 )

[

(

∆)

tσ √n

(11)

Here, at the 5% significance level, the probability is 95% and, thus, the t value can be obtained from the relevant table as t ) 2.571 for ν ) 5, where n is the number of the measurements and ν ) n 1. In the present study, each measurement has been repeated 6 times, and thus, n is equal to 6 and ν is equal to 5.18,19 The estimated uncertainties in the values of the torque are in the range of 0.02-1.3%. Error analyses for derived quantities, such as engine-brake effective power, total fuel consumption, brakespecific fuel consumption, and brake-effective efficiency, have also been performed. The error analysis by applying the method of Kline and McClintrok18 indicates that the uncertainties of brake-effective power and brake mean effective pressure are in the range of 0.02-1.3%, and it should be clearly noted that the estimated errors in the basic measurements and the uncertainties of total fuel consumption, brake-specific fuel consumption, and brake-effective efficiency are in the range of 0.2-2.8%. Thus, the estimated errors in the measurements of the basic and derived quantities do not significantly influence the overall uncertainty of the final results. Details of the error analysis can be found in the Supporting Information.

3. Theoretical Study

xdFd + xgFg Fmix ) xd + xg

( ( (

C1 and C2, F1 and F2, and be are euro/L, kg/m3, and kg kW-1 h-1, respectively. The main properties of diesel fuel and gasoline are given in Table 1. 2.3. Error Analysis. In the present study, error analysis of the experimental data has been performed using Kline and McClintrok’s method18 given by Durgun.19 As previously explained, each measurement has been repeated 6 times and, thus, Student’s t distribution can be applied to estimate the confidence intervals using the following relation:

) ]

x1 + x2r2 ∆be 1+ - 1 100 x1 + x2s2 be

(10)

r1 ) 1, r2 ) C2/C1 ) 1.229, s1 ) 1, s2 ) F2/F1 ) 0.907 where C1 is cost of diesel fuel, C2 is cost of gasoline, F1 and F2 are densities of diesel fuel and gasoline, respectively, and ∆be/be is the variation ratio of brake-specific fuel consumption. The units of

3.1. Description of the Model for NDF. In the present model, a multizone thermodynamic-based model developed by Shahed25,26 and then Ottikkutti27,28 has been used and modified with new assumptions to calculate the complete engine cycle and engine characteristics. Detailed information about this model has been given in the previous studies of the author.11,12,14,29 Here, a brief description of the model has been presented. In this model, the spray injected into the combustion chamber is divided into several zones. The boundaries of these zones are determined from lines of constant equivalence ratios. Applying the first law of thermodynamics, ideal gas equation, and other basic relations to the cylinder charge (these zones), a system of ordinary differential equations for cylinder pressure and zone volumes have been obtained, and by solving these ordinary (22) Ferguson, C. R. Applied thermosciences. Internal Combustion Engines; John Wiley and Sons: New York, 1986. (23) Heywood, J. B. Internal Combustion Engine Fundamentals; McGraw-Hill: New York, 1989. (24) Bosch, R. AutomotiVe Handbook, 5th ed.; Robert Bosch GmbH: Stuttgart, Germany, 2000. (25) Shahed, S. M.; Flynn, P. F.; Lyn, W. T. A model for the formation of emissions in a direct-injection diesel engine. Combustion Modeling in Reciprocating Engines; Mattavi, J. N., Amann, C. A., Eds.; Plenum Press: New York, 1980; pp 345-368. (26) Shahed, S. M.; Chiu, W. S.; Lyn, W. T. A mathematical model of diesel combustion. Proc. Inst. Mech. Eng. 1975, C94/75, 119–128. (27) Ottikkutti, P. Multizone modeling of a fumigated diesel engine. Ph.D. Thesis, Iowa State University of Science and Technology, Ames, IA, 1989. (28) Ottikkutti, P.; Gerpen, J. V.; Cui, K. R. Multizone modeling of a fumigated diesel engine. Society of Automotive Engineers (SAE) Tech. Pap. 910076, 1991. (29) S¸ahin, Z.; Durgun, O. Multi-zone combustion modeling for the prediction of diesel engine cycles and engine performance parameters. Appl. Therm. Eng. 2008, in press, ATE-2007-447R3.


differential equations simultaneously during the engine cycle using the Runge-Kutta 4 method, cylinder pressure and zone volumes can be calculated. Also, the temperatures of the zones can be computed from the ideal gas equation using cylinder pressure and zone volumes. In the present model, Dent’s30 correlation for spray penetration and Reitz’s correlation given by Heywood23 and Wakuri’s31 correlation for spray angle have been used. The effect of swirl on spray penetration and cone angle has been taken into account using Hiroyasu’s32 approach. For determination of instantaneous total mass and instantaneous mass rates of spray zones, it is required to know the spatial distribution of diesel fuel in the zones. In the present study, the concentration distribution of the fuel along the spray axis has been assumed to be hyperbolic, while across the spray, it has been taken as a normal distribution curve by benefit of information given in the literature.25-28 In this model, thermodynamic properties and their partial derivatives have been computed for the unburned mixture and the burned equilibrium products using Olikara’s method.33 The instantaneous total heat transfer from the cylinder contents to the combustion chamber walls has been calculated using Annand’s34 correlation. During the compression process and the expansion process, differential equations for cylinder pressure and temperature given by Heywood have been used to determine cylinder pressure and temperature values. Also, intake and exhaust processes have been computed by Durgun’s semi-empirical method.35,36 In the present study, residual gases, which remain in the cylinder from the previous cycle, have been taken into account using a coefficient of residual gases at the intake process calculations because they have an important effect on the intake gas properties. However, the effects of the residual gases were neglected in most of the theoretical engine cycle studies. In the present model, residual gas temperature has been chosen approximately at the beginning of cycle calculations. Then, after the cycle calculations were completed, chosen and calculated exhaust temperatures were compared. If the difference between these values is higher than 2%, the final value has been taken as the exhaust temperature and the cycle calculations have been repeated again. This calculation procedure has been applied iteratively until the difference between these values becomes smaller than 2%. Thus, complete cycle control has been performed. In some theoretical cycle models, a correction factor of indicator diagram φi has been used to take into account injection advance, ignition delay, and valve timing effects. The numerical values of the correction factor φi were given as 0.92-0.95. However, in the present model, injection advance and ignition delay have been considered and the pressure-volume (30) Dent, J. C. A basis for comparison of various experimental methods for studying spray penetration. Society of Automotive Engineers (SAE) Tech. Pap. 710571, 1971. (31) Wakuri, Y.; Fujii, M.; Amitani, T.; Tsuneya, R. Studies on the penetration of fuel spray in a diesel engine. Bull. Jap. Soc. Mech. Eng. 1960, 9, 123–130. (32) Hiroyasu, H.; Kadota, T.; Arai, M. Fuel spray characterization in diesel engines. Combustion Modeling in Reciprocating Engines; Mattavi, J. N., Amann, C. A., Eds.; Plenum Press: New York, 1980; pp 369404. (33) Olikara, C.; Borman, G. L. A computer program for calculating properties of equilibrium combustion products with some applications to I.C. engines. Society of Automotive Engineers (SAE) Tech. Pap. 750468, 1975. (34) Annand, W. J. D. Heat transfer in the cylinders of reciprocating internal combustion engines. Proc. Inst. Mech. Eng. 1963, 177, 973–990. (35) Durgun, O. A practical method for calculation engine cycles. Union of Chambers of Turkish Engineers and Architects, Chamber of Mechanical Engineer, 1991; Vol. 383, pp 18-29 (in Turkish). (36) Durgun, O. Internal combustion engines. Lecture notes, Karadeniz Technical University, Trabzon, Turkey, 1992 (in Turkish).

Energy & Fuels, Vol. 22, No. 5, 2008 3205

variation during the combustion process has been determined at a rounded character. Here, only the exhaust valve opening advance must be taken into account. For this reason, φi was chosen somewhat higher than the usual values. Here, φi has been selected as 0.98.29,35,36 Thus, indicated work and indicated efficiency obtained from gross cycle simulation have been corrected as follows: ηi )

Wgrφi(1 + Fstφ) Wi(1 + Fstφ) ) Fstφ(1 - f)QLHV Fstφ(1 - f)QLHV

(12)

where Wgr is the work performed per unit mass of cylinder contents. Wi is indicted useful work, Fst is the stoichiometric fuel-air ratio, φ is the equivalence ratio, QLHV is the lower heating value of the fuel, and f is residual gases mass fraction. In the present paper, hmin is equal to 1/Fst. After the complete diesel engine cycle was determined, engine performance parameters, such as brake-effective power, brakeeffective efficiency, brake-specific fuel consumption, etc., were calculated using relationships given by Heywood23 and Durgun.35,36 The present model can also be used for determining exhaust emissions. Here, mole fractions of CO and H2 have been computed on the basis of chemical equilibrium, and NO concentration has been calculated using the three coupled reactions of the extended Zeldovich mechanism. Detailed information about this model, calculation method of exhaust gases, and a flowchart of the model have been given in the previous studies of the author.11,12,14,29 3.2. Modifications in the Developed Model for Gasoline Blends. In the present study, the diesel engine cycle model for NDF is adapted to gasoline blends as follows. In the gasoline blend applications, gasoline and diesel fuel are mixed in the fuel tank and injected using the usual injector. The properties of these mixtures have been calculated using the above-mentioned eqs 7-9. Also, the gas constant R and the CN of the mixture have been computed using the following formulas:8 Rmix )

xdFdRd + xgFgRg xdFd + xgFg

CNmix )

(13)

xdCNd + xgCNg xd + xg

(14)

Because any empirical equation about spray penetration and spray angle calculations for gasoline mixtures has not been given in the literature yet, the empirical equations developed for NDF have been used to calculate spray penetration and spray angle in this study. To obtain the properties of the combustion products for any blend consisting of two fuels, the following combustion reaction could be written:

[

y13

[

]

(xdC13.342H24.74O0.0049 + xgC7H17) f c + 0.25h - 0.5oy (O2 + 3.7274N2 + 0.0444Ar) + φ (15)

y1H + y2O + y3N + y4H2 + y5OH + y6CO + y7NO+ y8O2 + y9H2O + y10CO2 + y11N2 + y12Ar

]

where xd and xg are the volumetric percentages of diesel fuel and gasoline, respectively. Here, the mole numbers of all of the products and 12 unknown mole fractions yi have been determined using Olikara’s method.33 After these mole numbers and mole fractions were calculated, thermodynamic properties and their partial derivatives, depending upon temperature,


S¸ahin

Figure 3. Comparison of pressure values obtained from the present model and experiments of Rakopoulos et al.2 as functions of gasoline crank angle.

Figure 2. (a and b) Comparison of brake-effective power and brakeeffective efficiency values obtained from the present model and the experimental results of Bilgin et al.5 as functions of ethanol ratio %, respectively. Table 2. Comparison of the Obtained Results from the Present Model with the Ethanol Experiments of Bilgin et al.5 ε ) 21, ethanol (%)

∆Ne/Ne (%)

∆ηe/ηe (%)

0 2 4 6

11.691 11.190 9.545 11.767

11.661 14.235 12.892 13.194

pressure, and equivalence ratio, have been computed using functions given by Ferguson.22 3.3. Accuracy Control of the Presented Model for Gasoline Mixtures. In the present study, accuracy controls of the developed model have been performed for NDF, gasoline and ethanol fumigation, gasoline and ethanol blends, and heat balance. For the cases of NDF and fumigation, detailed comparisons to the experimental and other relevant theoretical results have also been performed in the previous studies of the author.11,12,14,29 A satisfactory conformity can be observed between these results. Here, some examples of these comparison results for ethanol and gasoline blends have been presented. In parts a and b of Figure 2, Table 2, and Figure 3, brake-effective power, brakeeffective efficiency, and cylinder pressure obtained from the present model have been compared to experimental results for ethanol-diesel fuel blends of Bilgin et al.5 and Rakopoulos et al.,2 respectively. Comparisons of brake-effective efficiency and brake-effective power obtained from the present model and experimental results for gasoline blends have also been shown

Figure 4. (a and b) Comparison of brake-effective power and brakeeffective efficiency values obtained from the present model and experimental results as functions of gasoline ratio %, respectively.

in parts a and b of Figure 4. A satisfactory conformity can be observed in Figures 2-4 and Table 2 for ethanol and gasoline blends. Experiments presented in this study for gasoline blends and experiments of Bilgin et al. for ethanol blends have been performed in a single-cylinder diesel engine at Karadeniz Technical University, Engineering Faculty, Mechanical Engineering Department, Internal Combustion Engines Laboratory. For this test engine, at a few instances, the maximum difference ratios of the performance parameters reached levels of 18%. This could be attributed to approximation of some of the parameters of the test engine that were not measured and selecting approximately these parameters, aging of the exhaust temperature thermocouple, and insufficiency of the spray penetration correlation. For example, when Dent’s spray penetration correlation was used, most of the injected fuel was predicted to burn toward the end of the expansion period.



However, when Hiroyasu’s correlation was used, predicted combustion temperatures were as high as 4000 K. For this reason, the spray penetration correlation was arranged by averaging Dent’s and Hiroyasu’s correlations after a lot of numerical applications, and the following correlation has been adapted for the test engine: xt ) 0.65xt,D + 0.35xt,H

(16)

In the test engine, equivalence ratios range between 1.086 and 1.149. However, equivalence ratios of vehicle diesel engines are generally in the range of 0.650-0.80, and various combustion approximations and parameters, which have been developed for vehicle diesel engines, are valid for these equivalence ratios. Thus, in the computations, more fuel exists in the engine cylinder than for the test engine. Then, as more fuel burns in the last burning zone, calculated combustion and exhaust temperatures are higher than experimental results. However, as it can be seen from parts a and b of Figure 2, Figure 3, and parts a and b of Figure 4, predicted values agree reasonably with the measured data generally. On the other hand, for determining the mechanical losses and mechanical efficiency, an empirical mean pressure of mechanical losses formula developed for vehicle diesel engines has been used. This equation may not predict satisfactorily mechanical losses of the test engine. A comparison has been performed with Rakopoulos’s experimental results for a mixture containing 15% ethanol. As shown in Figure 3, the calculated cylinder pressure values from the model are lower than Rakopoulos’s experimental values during the combustion process, but cylinder pressure values are higher than Rakopoulos’s2 values during the expansion process. This difference might have arisen from spray penetration equations used in the present study. A different equation for spray penetration for ethanol blends is required, or a correction factor has to be added to the spray penetration equation developed for NDF. However, any spray penetration equation for ethanol blends has not been found in the literature because theoretical studies on ethanol blends are very scarce. Nevertheless, the brake-effective efficiency value obtained from the present model is about at the levels of 0.3517 and the brakeeffective efficiency given by Rakopoulos et al. is 0.3288 for 15% ethanol. Thus, the difference in brake-effective efficiencies between the present model prediction and the experimental result of Rakopoulos et al. is 6.96%. Also, at the same ethanol ratio, exhaust temperature values obtained from the present model and given by Rakopoulos et al. are 608 and 660 K, respectively. Hence, the difference between these temperatures is 7.88%. 5. Results and Discussion 5.1. Results of the Experimental Study. In the present study, the effects of 2-10% gasoline blends have been investigated experimentally at the speeds of 900-1600 rpm at 1:1 gas position (at wide-open throttle) for selected compression ratios of 18-23. Here, 216 experiments have been performed for seven engine speeds, six compression ratios, and six gasoline blends ratios. Also, all of the experiments have been repeated 6 times, and the mean values of these measurements have been used in computations to minimize the measurement errors. Thus, totally 1296 experiments have been carried out in this study. Some examples of the experimental results have been shown in Figures 5-7 and Tables 3-8. Variation ratios of brakeeffective power, brake-specific fuel consumption, brake-effective efficiency, and cost have also been given as various tables for selected engine speeds at all compression ratios. Also, variations

Figure 5. (a-c) Variations of brake-effective power as functions of the gasoline ratio. (d) Variations of brake-effective power as functions of the engine.

of brake-effective power and brake-specific fuel consumption as functions of the compression ratio have been presented in parts a-d of Figure 8.


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Figure 6. (a-c) Variations of brake-specific fuel consumption as functions of the gasoline ratio.

Figure 7. (a-c) Variations of brake-effective efficiency as functions of the gasoline ratio %.

It can be seen in parts a-d of Figure 5 that brake-effective power decreases generally at low compression ratios and selected engine speeds as the gasoline ratio increases. However, at high compression ratios, brake-effective power increases generally up to 6-8% gasoline-diesel fuel blends. The experimentally obtained increases in the ratios of brake-effective power are approximately at the levels of 2-4%. It is well-known that the heating value of the fuel affects the engine power. The LHV of gasoline is somewhat higher than that of diesel fuel. Thus, the higher energy density (capacity) of the gasoline-diesel fuel blends cause a little increment in the engine power when used in the test diesel engine without any modifications. However, the higher LHV is not the only factor that results in the power increment in the present experiments. Some other factors could affect brake-effective power. For example, CN and density of gasoline are lower than that of diesel fuel. Therefore, as the gasoline ratio increases, the CN and density of the mixture decrease. Thus, because of reduction of CN of the mixture, ignition delay may increase slightly. The longer ignition delay could result in an inefficient heat-work conversion process. This inefficient combustion results in more reduction in the power. In addition, the lower viscosity of the

gasoline-diesel blends leads to greater pump and injector leakages, reducing fuel delivery and hence power output decreases.6,7 Brake-specific fuel consumption decreases at low gasoline ratios, but it increases at high gasoline ratios as shown in parts a-c of Figure 6. The decrease in the variation ratios of brakespecific fuel consumption are in the range of 3-6%. As explained above, gasoline has a higher LHV than diesel fuel. By increasing the gasoline ratio, LHV of the mixture increases. Thus, brake-effective power increases and brake-specific fuel consumption decreases at high gasoline ratios. The self-ignition temperature of diesel fuel is lower than that of gasoline (Table 1). Thus, in the injected spray of the blends, the diesel fuel could initiate ignition. Because the boiling point of gasoline is lower than that of diesel fuel, gasoline will evaporate before diesel fuel, and after the first auto-ignition of diesel fuel, fully evaporated gasoline could burn more rapidly than diesel fuel. Thus, swirl and additional gas motions take place in the cylinder because accumulated gasoline burns more rapidly than diesel fuel. These swirl and gas motions would mix rather well the gasoline-diesel fuel and air mixture through the unburned blend spray. As a result, the diesel engine



Table 3. Variation Ratios of Brake-Effective Power, Brake-Specific Fuel Consumption, Brake-Effective Efficiency, and Cost n ) 1300 (rpm)

ε ) 18

ε ) 19

gasoline (%)

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

2 4 6 8 10

-7.692 -7.085 -4.251 -5.263 -4.049

6.192 1.858 1.858 2.167 0.310

-7.491 -3.745 -4.200 -4.200 -2.247

6.575 2.594 2.964 3.649 2.133

2.529 -1.379 -0.120 1.996 0.0

-4.403 -3.775 -3.145 -3.459 -5.975

4.494 3.745 2.621 3.371 5.618

-4.058 -3.079 -2.035 -2.058 -4.267


ε ) 20

ε ) 21

gasoline (%)

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

2 4 6 8 10

5.870 1.619 3.036 -0.405 1.417

-6.502 -6.192 -6.192 -3.096 -7.210

6.464 6.084 6.084 2.662 6.844

-6.165 -5.514 -5.173 -1.690 -5.524

0.388 0.777 1.359 0.388 -0.194

-0.658 -3.289 -0.968 -0.968 -0.968

0.358 2.867 0.716 0.358 0.358

-0.300 -2.590 0.108 0.469 0.831


ε ) 22

ε ) 23

gasoline (%)

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

2 4 6 8 10

1.754 2.534 1.754 2.144 0.585

-4.127 -7.937 -6.349 -7.302 -5.715

4.074 8.148 5.926 7.037 5.556

-3.781 -7.272 -5.332 -5.957 -4.001

0.588 2.157 0.0 2.549 1.961

-3.175 -6.032 -3.492 -6.984 -6.984

3.346 5.948 3.346 7.435 7.063

-2.826 -5.352 -2.444 -5.635 -5.350


ε ) 18

ε ) 19

gasoline (%)

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

2 4 6 8 10

-3.065 -5.567 -3.532 -3.532 -2.974

4.00 6.462 3.692 3.077 1.538

-3.984 -6.513 -3.984 -3.448 -1.916

4.375 7.231 4.818 4.572 3.383

1.612 -0.719 -1.617 0.540 0.0

-3.406 -4.334 -1.548 -2.477 -4.954

3.041 4.182 0.760 1.901 4.563

-3.057 -3.643 0.479 -1.062 -3.227


ε ) 20

ε ) 21

gasoline (%)

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

2 4 6 8 10

4.152 0.0 1.083 0.361 1.083

-4.025 -3.715 -3.406 -3.096 -5.573

4.183 3.422 3.042 2.662 5.323

-3.679 -3.019 -2.357 -1.690 -3.857

3.041 3.936 3.578 3.578 1.252

-2.532 -5.063 -3.481 -4.114 -2.532

2.612 5.224 3.358 3.731 2.239

-2.180 -4.377 -2.433 -2.723 -0.761


ε ) 22

ε ) 23

gasoline (%)

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

∆Ne/Ne

∆be/be

∆ηe/ηe

∆C/C

2 4 6 8 10

1.389 1.389 0.0 2.951 -0.694

0.0 -2.632 -0.329 -3.618 -0.658

0.0 2.509 0.0 5.584 0.358

0.361 -1.928 1.086 -2.220 1.147

1.579 2.632 0.526 2.632 1.579

-0.649 -2.922 -0.654 -4.545 -3.571

0.727 2.909 0.364 4.364 3.636

-0.291 -2.220 0.425 -3.160 -1.819

combustion process could be improved because of faster and more efficient burning of the fuel. As shown in parts a-c of Figure 7, generally similar improvements have been obtained for brake-effective efficiency because LHVs of diesel fuel and gasoline are close to each other. In the present study, the effects of engine speed and compression ratio for various gasoline blends have also been studied, and it is determined that gasoline blends give better results for engine performance at high engine speeds and high compression ratios than that of low engine speeds and low compression ratios for this engine. Because the available time for combustion of diesel fuel decreases at high engine speeds, faster mixing rates of diesel fuel with air are required at high engine speeds. For gasoline blends, this could be achieved by

faster burning of gasoline and additional gas motions.11-14 It is well-known that diesel combustion is largely controlled by the mixing rate of diesel fuel and air.23 It is thought that the mixture of the residual gas and the air is compressed to the higher pressures and temperatures at high compression ratios, and then the gasoline-diesel fuel blend is injected in this mixture. In this case, the injected gasoline-diesel fuel blend evaporates more efficiently, and thus, a homogeneous mixture could be formed than that of lower compression ratios. On the other hand, a little increment in ignition delay occurs at low gasoline blend ratios, and thus, more fuel burns close to the TDC and brakeeffective efficiency, and brake-effective power would increase. At gasoline ratios higher than about 10% (which were not tested


S¸ahin

Figure 8. (a-d) Variation ratios of brake-effective power and brake-specific fuel consumption as functions of ε, respectively.

Figure 10. (a and b) Predicted variations of temperature and pressure as functions of the gasoline ratio %, respectively. Figure 9. (a and b) Predicted variation of ignition delay and equivalence ratio as functions of the gasoline ratio %, respectively.

in this study), ignition delay would increase significantly and engine knock could occur. In parts a-d of Figure 8, variation ratios of brake-effective

power and brake-specific fuel consumption have been shown. The most favorable percentage of gasoline is 4-6% for this engine at the selected compression ratios. As shown in Tables 3-8, gasoline blends are not economical because the cost of gasoline is higher than that of diesel fuel in Turkey, as well as in many other countries. However, approximately a 3% cost reduction is attained at low gasoline ratios


Figure 11. (a and b) Predicted variations of NO concentration and mole fractions of CO and H2 as functions of the gasoline ratio %, respectively.

for high compression ratios. Generally, low prices have been obtained at low gasoline ratios, such as 4-6%, at selected compression ratios. In the present study, the effects of 2-10% gasoline blends have been investigated experimentally at the speeds of 900-1600 rpm at 1:1 gas position (at wide-open throttle) for selected compression ratios of 18-23. The experimentally obtained results are valid for the experimental test engine and fuel prices in Turkey. To obtain generalized results, a systematic set of experiments must be performed for various vehicle diesel engines. The cost analysis relation given here can be applied for fuel prices in any country. 5.2. Theoretical Results. Here, the effects of 2-10% gasoline blends on the cylinder gas pressure, cylinder gas temperature, ignition delay, equivalence ratio, and exhaust emissions have been investigated theoretically using the multizone thermodynamic model. Engine characteristics, such as brake-effective power, brake-effective efficiency, and brakespecific fuel consumption, exhibit a similar tendency to that of experimental results. For this reason, these theoretical results have not been repeated here again. In subsequent studies, it is planned to measure exhaust emissions and determine the p-V diagram experimentally. The necessary equipment has been installed, and detailed experimental studies will be performed in the near future. Here, some combustion characteristics and exhaust emissions determined theoretically have been presented. In the theoretical studies, selected engine speeds are 1300 and 1500 rpm, and selected compression ratios are 19, 21, and 23. Some of the obtained results from the theoretical model have been presented in Figures 9-11 and Table 9. As shown in Figure 9a, the predicted ignition delay increases with the increasing percentage of gasoline in the mixture because


CN of the gasoline-diesel fuel mixture is lower than NDF. It is well-known that diesel fuel has a high CN at the levels of 45-50, but the CN of gasoline is significantly lower. In the present study, CN of gasoline is taken as 3.11,22-24 Lower CN means longer ignition delay, allowing more time for fuel to vaporize before combustion starts. Thus, longer ignition delay results in a high combustion temperature, which causes a rapid heat release. Also, as the gasoline percentage in the blend increases, the equivalence ratio decreases slightly as shown in Figure 9b. Predicted combustion duration remains constant for all selected compression ratios at 1300 rpm, but it decreases at 1500 rpm for high gasoline ratios. After first burning (known as the rapid burning period) of the mixture,23 the injected fuel blends mix with air homogenously and burn very rapidly because of swirl and additional gas motions, which is formed as a result of the fast gasoline combustion. Thus, combustion duration decreases. As combustion duration decreases, the same amount of fuel burns in less time and, thus, as shown in Figure 10a, the temperature of the cylinder charge increases. However, at higher temperatures, the formation rate of exhaust emissions increases. As shown in Figure 10b, there is no significant difference between the predicted cylinder pressures for gasoline blends and NDF. However, the combustion temperature and also exhaust temperature increase with an increasing gasoline percentage in the blends. However, generally, for greater than after 8-10% gasoline ratios, the incremental ratios of temperature become smaller, because the combustion temperature and exhaust temperature increase, as shown in parts a and b of Figure 11. The NO concentration and mole fractions of CO and H2 increase too. The mechanism of NO formation is greatly dependent upon temperature because of the high activation energy needed for the reactions involved.37,38 Gasoline combustion normally produces higher temperatures than diesel combustion because of its slightly higher heating value and its higher burning speed. Also, lower CN means a change in the ignition delay and more accumulated fuel/air mixture, which cause a steep heat release at the beginning of the combustion, resulting in higher temperatures and higher NO formation. In the present model, the burning speed of gasoline has not been taken into account. For this reason, it is thought that the whole effect of gasoline in the blends could not be reflected in the results. Ignition delay increases with an increasing gasoline ratio. This could cause higher temperatures throughout the cycle, especially during the combustion period as shown in Figure 10a. Thus, the formation of CO increases because a higher combustion temperature significantly affects its formation, that is, because higher temperatures increase the dissociation rate of CO2 and, thus, the formation rate of CO increases.3,37 Also, it is well-known that the carbon monoxide ratio from internal combustion engines could be controlled primarily by the air/fuel equivalence ratio and it decreases as equivalence ratio increases.37,38 The equivalence ratio decreases slightly with increasing gasoline percentage in the mixture, and thus, the mole fraction of CO increases. On the other hand, a thickened quench layer created by the cooling effect of vaporizing gasoline could play a major role in the increased CO production.3 As the temperature of the cylinder charge increases, the (37) Sayın, C.; Uslu, K.; Çanakçı, M. Influence of injection timing on the exhaust emissions of a dual-fuel CI engine. Renewable Energy 2008, 33 (6), 1314–1323. (38) Jiang, Q.; Ottikkutti, P.; Gerpen, J. V.; Meter, D. V. The effects of alcohol fumigation on diesel flame temperature and emissions. Society of Automotive Engineers (SAE) Tech. Pap. 900386, 1990.


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Table 9. Predicted Variation Ratios of NO Concentration and Mole Fractions of CO n ) 1300 (rpm)

ε ) 19

ε ) 21

ε ) 23

gasoline (%)

∆NO/NO (%)

∆CO/CO (%)

∆NO/NO (%)

∆CO/CO (%)

∆NO/NO (%)

∆CO/CO (%)

2 4 6 8 10

-0.119 1.835 5.599 7.601 2.431

21.312 30.243 37.618 86.843 57.031

3.736 7.958 9.109 10.019 -3.992

30.623 41.389 35.035 15.679 6.828

1.225 9.155 13.414 16.111 15.464

22.531 37.408 51.988 63.116 67.193

formation of H2 increases, because the H2O dissociation rate increases at high temperature. NO and CO emissions increase for gasoline blends. To decrease these emissions, optimization of the operation of the test engine must be preformed. For example, injection advance, equivalence ratios, and injection pressure may be optimized for gasoline blends. 6. Conclusions The main results and recommendations achieved from the present study can be summarized as follows: (1) From experimental results, it is determined that gasoline blend ratios decrease brake-effective power output at the levels of 1.5-4% at low compression ratios. However, brake-effective power increases 1.5-4% at high compression ratios. (2) Brake-effective efficiency increases at the levels of 2-6% and brake-specific fuel consumption decreases nearly 2-6% at low gasoline blend ratios. Because the cost of gasoline is higher than that of diesel fuel and the decrease in ratios of brake-specific fuel consumption are low, gasoline blends are not economical for the test engine at low compression ratios. However, approximately, a 3% price reduction is attained at low blend ratios for high compression ratios. (3) From theoretical results, it is determined that ignition delay increases as gasoline ratio increases, because of decrement of CN of the mixture. Combustion duration remains approximately constant for gasoline blends at low engine speed, but it decreases at high engine speed. The mole fraction of CO and NO concentrations increase with increasing gasoline percentages in the blends because of high combustion temperatures. As gasoline addition increases, NO concentration and CO mole fraction predicted for the test engine increase approximately 7-29 and 15-90%, respectively. (4) As a result, it can be said that 4-6% gasoline blends can be applied in the test engine without any modification. Thus, brake-effective power of the engine could increase at the levels of 2% and brake-specific fuel consumption could decrease nearly 3%. Thus, a 3% fuel cost decrement could be attained. Generally, exhaust emissions increase with gasoline blends. In Turkey, some truck drivers add a little gasoline to diesel fuel because of their experience of increased engine performance when they drive long distances. Similar results might be attained for vehicle diesel engines. However, to determine the most favorable blend ratio for any vehicle diesel engine, to achieve general results, and to give general recommendations, more systematic experimental and theoretical studies for actual vehicle diesel engines must be performed. (5) NO and CO emissions increase with increasing gasoline blends. To decrease these emissions, optimization of the operation parameters of various engines must be performed. For example, injection advance or injection pressure may be optimized for gasoline blends. Also, gasoline blends for constant equivalence ratios may be investigated for different engines. The author has investigated gasoline and

ethanol fumigation for constant equivalence ratios in the earlier studies. Constant equivalence ratio applications give good results for NO and CO emissions than variable equivalence ratios. Thus, it would be better to investigate gasoline blends at constant equivalence ratios. (6) It is planned to perform more systematic experimental studies on the use of gasoline-diesel fuel blends in actual multicylinder vehicle DI diesel engines in the near feature. Fundamental studies for this purpose have been started, which are supported by State Planning Institutes of Turkey. In these studies, it is planned to use comprehensive exhaust emissions measuring devices and determine the p-V diagram experimentally. Acknowledgment. The author expresses her gratitude to Prof. Dr. Orhan Durgun for his very valuable comments during the preparation and process of this manuscript.

Nomenclature be ) brake-specific fuel consumption (kg kW-1 h-1) Bt ) total (hourly) fuel consumption (kg/h) C ) cost of the fuel (euro) Ne ) effective power (kW) n ) engine speed (rpm) QLHV ) lower heating value (kJ/kg) R ) gas constant (kJ kg-1 K-1) Tb ) brake torque (N m) Greek Symbols F ) density (kg/m3) ε ) compression ratio φ ) equivalence ratio ηe ) effective efficiency xg and xd ) volumetric percentage of gasoline and diesel fuel, respectively xhum ) humidity correction factor Subscripts d ) diesel g ) gasoline hum ) humidity mix ) mixture AbbreViations CO ) carbon monoxide CN ) cetane number DI ) direct injection LHV ) lower heating value NDF ) neat diesel fuel NO ) nitric oxide TDC ) top dead center Supporting Information Available: More detailed error analysis. This material is available free of charge via the Internet at http://pubs.acs.org. EF800236Y

Experimental and Theoretical Investigation of the Effects of Gasoline

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