Master of Engineering

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1 Ethanol Fuel as Renewal Energy in Spark Ignition Engines Julio César Villavicencio Cevallos Master of Engineering Manufacturing Engineering and Management 2015

2 Table of contents Table of contents... i List of Figures... ii List of Tables... v Appendix... v Abstract... vi 1. Introduction Literature Review Specific Heat Ratio and Compression Ratio in Engine Thermal Efficiency using Ethanol Ethanol/ Gasoline Energy Radio Mass Fraction Burn Effect of Ethanol on Gas Emissions Experimental Procedure Results and Discussion Performance of SI Engine with Ethanol Direct Injection (EDI) and Gasoline Port Injection (PFI) p-v Diagrams Indicated Mean Effective Pressure Brake Mean Effective Pressure Volumetric Efficiency Brake Specific Fuel Consumption Brake Specific Emissions Ethanol Effect on Burning duration, Temperature Drop, and Compression Ratio i

3 4.2.1 Mass Fraction Burn and Ethanol Energy Ratio Ignition Time Delay Combustion Time Period Temperature Drop at the End of Combustion Process Estimation of Compression Ratio Conclusions Appendix References & Bibliography List of Figures Figure 1. Brake thermal efficiency at compression ratio of 15 [7]... 3 Figure 2. Spark sweep at 2000 rpm, 8 bar BMEP [8]... 4 Figure 3. MFB Curves for gasoline, E50, and E85. Single and split injection [19]... 8 Figure 4. Flame speed and MFB (IMEP 5 bar) [18]... 8 Figure 5. Flame images of Bio-fuels under dual-injection strategy (IMEP 5 bar) [18]... 9 Figure 6. Mass fraction burned for PFI and DI operation [20] Figure 7. Cumulative HC emission for blends E5 to E85 at 22 C [21] Figure 8. Effect of unleaded gasoline-ethanol blends on HC emissions at different compression ratios [23] Figure 9. CO concentration versus BMEP at 20 BTDC [10] Figure 10. Effects of lambda on NOx emissions versus spark timing [26] Figure 11. Comparison of NOx emissions for gasoline and ethanol [26] Figure 12. NOx emissions vs. speed at a constant load of 340 kpa [27] Figure 13. NOx emissions vs. speed at a constant load of 510 kpa [27] ii

4 Figure 14. Effect of load on NOx emissions at a constant speed of 2000 rpm [27] Figure 15. Schematic of the engine system [11] Figure 16. p-v diagrams: Different ethanol 3500 rpm Figure 17. p-v diagrams: Different ethanol 4000 rpm Figure 18. p-v diagrams: Different ethanol 4500 rpm Figure 19. p-v diagrams: Different ethanol 5000 rpm Figure 20. IMEP vs. ethanol 3500 rpm Figure 21. IMEP, Torque and Fuel content at 3500 rpm Figure 22. IMEP vs. ethanol 4000 rpm Figure 23. IMEP and torque at 4000 rpm Figure 24. IMEP vs rpm Figure 25. IMEP and torque at 4500 rpm Figure 26. IMEP and torque at 4500 rpm Figure 27. IMEP vs. ethanol 5000 rpm Figure 28. Summary IMEP vs. Ethanol content at different RPM Figure 29. Variation of BMEP with EER [11] Figure 30. Zoom in on Figure 29 - Medium load Figure 31. BMEP vs. EER Medium load using Equation Figure 32. Variation of Volumetric efficiency with EER [11] Figure 33. Zoom in on Figure 32 - Medium load Figure 34. Volumetric efficiency vs. EER - Medium load using Equation Figure 35. Volumetric efficiency vs. EER - Medium using Equation 7 and real RPM Figure 36. Variation of BSFC with EER [11] Figure 37. Zoom in on Figure 36 - Medium load Figure 38. BSFC vs. EER - Medium load using Equation Figure 39. BSFC vs. EER Medium load using Equation 8 and real RPM.. 33 Figure 40. Variation of BSCO with EER [11] Figure 41. BSCO vs. EER Medium load Figure 42. Variation of BSHC with EER [11] iii

5 Figure 43. BSHC vs. EER Medium load Figure 44. Variation of BSNO with EER [11] Figure 45. BSNO vs. EER Medium load Figure 46. Log p vs. Log v; 3500 rpm; G29 E Figure 47. MFB curves at 3500 rpm Figure 48. Crank angle variation vs. EER at 3500 rpm: Time delay Figure 49. Crank angle variation vs. EER at 3500 rpm: Combustion period. 40 Figure 50. MFB curves at 4000 rpm Figure 51. Crank angle variation vs. EER at 4000 rpm: Time delay Figure 52. Crank angle variation vs. EER at 4000 rpm: Combustion period. 42 Figure 53. MFB curves at 4500 rpm Figure 54. Crank angle variation vs. EER at 4500 rpm: Time delay Figure 55. Crank angle variation vs. EER at 4500 rpm: Combustion period. 44 Figure 56. MFB curves at 5000 rpm Figure 57. Crank angle variation vs. EER at 5000 rpm: Time delay Figure 58. Crank angle variation vs. EER at 5000 rpm: Combustion period. 46 Figure 59. Mass fraction burned for PFI and DI operation [20] Figure 60. CAD (0-5%MFB) vs. RPM Figure 61. CAD (0-10%MFB) vs. RPM Figure 62. CAD (10-90%MFB) vs. RPM Figure 63. Representation of the constant-volume cycle in the p, v and T, s diagram [34] Figure 64. Temperature Drop 3500 rpm Figure 65. Temperature Drop 4000 rpm Figure 66. Temperature Drop 4500 rpm Figure 67. Temperature Drop 5000 rpm iv

6 List of Tables Table 1. Thermal efficiency of an SI engine for regular gasoline and E85 [6] 2 Table 2. Specification of the engine [11] Table 3. Combustion time delay and period at 3500 rpm Table 4. Combustion time delay and period at 4000 rpm Table 5. Combustion time delay and period at 4500 rpm Table 6. Combustion time delay and period at 5000 rpm Table 7. Theoretical temperature at start of combustion: 3500 rpm Table 8. Theoretical temperature at start of combustion: 4000 rpm Table 9. Theoretical temperature at start of combustion: 4500 rpm Table 10. Theoretical temperature at start of combustion: 5000 rpm Table 11. Temperature at End of Combustion 3500 rpm Table 12. Temperature at End of Combustion 4000 rpm Table 13. Temperature at End of Combustion 4500 rpm Table 14. Temperature at End of Combustion 5000 rpm Table 15. Estimated Compression Ratio: 3500 rpm Table 16. Estimated Compression Ratio: 4000 rpm Table 17. Estimated Compression Ratio: 4500 rpm Table 18. Estimated Compression Ratio: 5000 rpm Appendix Appendix 1. Calculation of IMEP using average pressure v

7 Abstract The outstanding features of ethanol fuel and its consumption in many countries as a renewal energy in spark ignition engines have attracted the attention of researchers all over the world. Accordingly, a former Ph. D student of the University of Technology Sydney collected medium load data in a single-cylinder engine. This engine was modified to work with ethanol direct injection (EDI) plus gasoline port injection (GPI). The base data gathered consists of pressure, crankshaft angles, torque, gas emissions, and fuel consumptions. These data was captured at four different engine speeds ranging from 3500 to 5000 rpm and different ethanol energy ratios (EER). The EERs vary from zero (gasoline only) to sixty-one (a mixture of gasoline and ethanol). The total energy of the fuel was held constant despite the different levels of EERs in the combustible. The data mention above was tabulated to evaluate several engine characteristics. The engine performance, pressure-volume diagrams, indicative mean effective pressure (IMEP), brake mean effective pressure (BMEP), volumetric efficiency, brake specific fuel consumption and gas emission were calculated, analyzed and discussed in this report. Similarly, the effect of ethanol on burning durations, mass fraction burn, ignition time delays and combustion time periods were included in this paper. Likewise, an estimation of the temperature drop and compression ratio was examined. The results indicated that EDI and GPI in spark ignition engines present remarkable characteristics. The use of ethanol in the fuel, for example, exhibited larger work in the pressure-volume diagram and significant falls in the in-cylinder temperature than gasoline fuel. Furthermore, parameters like IMEP, BMEP, and volumetric efficiency demonstrated improvements when ethanol was part of the fuel. Nevertheless, larger carbon monoxide and hydrocarbon emission than vi

8 gasoline were detected. The increase in gas emissions was attributed to the decrease in temperature at the end of the combustion. This temperature drop played a meaningful role in the emission of the HC and CO, but it decreased the emission of NOx due to the lower temperature reached. Finally, MFB curves were fundamental to broaden our understanding of ignition time delay and combustion time period. To illustrate, ignition time delays presented increased values when ethanol was part of the fuel. A lower ethanol heating value would be the reason for this behavior. Nevertheless, other ethanol quality like flame propagation reduced the combustion time period at every engine speed tested. vii

9 1. Introduction Ethanol or ethyl alcohol is a flammable and colorless liquid that possesses a low heating value, high octane number and heat of vaporization. Some of these characteristics have promoted its use as fuel for internal combustion engines in countries such as Brazil, Canada, Sweden, India, Australia, Thailand, China, Colombia, Peru, and Paraguay [1]. Particularly, blends of ethanol and gasoline containing between 10% (E10) to 85% (E85) of ethanol are used in flexible fuel vehicles (FFV) in Brazil [2]. These type of vehicles are designed to function either consuming gasoline, ethanol or a mixture of them. Seventy-eight percent of the global ethanol production is manufactured in Brazil and the United States. According to Gupta and Demirbas [1], the demand for ethanol would be more than the double in the coming years due to ethanol more remarkable advantages such as being less toxic than methanol, high octane number resulting in better thermal efficiency and engine power, higher compression ratios and antiknock effect when compared with gasoline [3]. Accordingly, ethanol s features and influence on spark-ignition (SI) engines make significant interest among researchers, so that more learning can be obtained about this topic. The aim of this report is to analyze and discuss the data obtained by a former Ph.D. student of the University of Technology Sydney. On top of that, the performance of a spark ignition engine using ethanol direct injection (EDI) and gasoline port injection (GPI) is evaluated. Additionally, the effect of ethanol on the polytropic index, burning duration, temperature drop, and compression ratio are investigated so that more knowledge can be acquired from the use of ethanol as a renewal energy in an SI engine with EDI and GPI. 1

10 2. Literature Review 2.1 Specific Heat Ratio and Compression Ratio in Engine Thermal Efficiency using Ethanol It is a fact that the increase of specific heat ratio and compression ratio enhance thermal efficiency. Actually, Shimada and Ishikawa [4] argue the use of hydrous ethanol (40 to 60 wt. % ethanol) to enrich the specific heat ratio of the operational gas by providing hydrogen in a lean combustion. In fact, hydrous ethanol increased thermal efficiency in 1.5 times that of conventional spark-ignition (SI) engines. Additionally, a previous work to investigate the effect of ethanol blends in thermal efficiency has been applied in an SI engine with three different compression ratios [5]. To illustrate, values of thermal efficiencies up to 2.8 points greater than those using regular gasoline have been tabulated at 9.2 compression ratio, wide open-throttled and 2500 rpm. Likewise, a compression ratio of 11.8 showed 4.5 point increase of thermal efficiency at 2500 rpm and 7.8 points greater for a compression ratio of and same speed. Table 1 displays a summary of these values. Table 1. Thermal efficiency of an SI engine for regular gasoline and E85 [6] 2500 rpm / WOT Compression Thermal efficiency ratio Regular gasoline E85 Δ

11 Though the increase of compression ratio can also augment knock effect, Nakama, Kusaka [7] affirm that the addition of ethanol suppresses this effect allowing the engine to work with high compression ratio. By doing so, the low calorific value of ethanol can be overwhelmed by increasing the compression ratio and adding ethanol. Figure 1 displays the effect of ethanol and high compression ratio in the thermal efficiency compared with a compression ratio of 9.5. Figure 1. Brake thermal efficiency at compression ratio of 15 [7] Similarly, high octane number of ethanol has been studied using the direct injection of E85 and port fuel injection of gasoline to reduce knock in the engine resulting in a high compression ratio [8]. The effect of ethanol usage is displayed in Figure 2 where thermal efficiency improves whether the percentage of E85 is increased. Finally, it can be said that there are multiple benefits of using ethanol as part of the fuel in an SI engine. For instance, some research has demonstrated the potential increase of thermal efficiency due to changes in the compression ratio and use of ethanol. Besides, the specific heat ratio could be enhanced if ethanol is 3

12 used as reported by Shimada and Ishikawa [4]. To sum up, thermal efficiency presents irresistible improvements thanks to ethanol properties. Figure 2. Spark sweep at 2000 rpm, 8 bar BMEP [8] 4

13 2.2 Ethanol/ Gasoline Energy Radio Ethanol possesses certain attributes that contribute to enhancing engine s performance. The improvement of volumetric efficiency, thermal efficiency as well as suppression of knocking effect are some of the effects of ethanol on the performance of spark ignition engines [9]. Compared with gasoline, CO, NOx and HC emissions are lower using ethanol [10]. In order to analyze the influence of ethanol on the engine performance, ethanol energy ratio and ethanol fraction energy have been used in multiple articles. Ethanol/Gasoline Energy Ratio (EER) is defined in Equation 1. Zhuang and Hong [11] emphasize this relation to analyzing brake mean effective pressure (BMEP), volumetric efficiency, brake specific fuel consumption (BSFC) and brake specific energy consumption (BSEC). In their investigation, EER was varied changing the mass flow of ethanol and gasoline, but keeping constant the denominator in Equation 1. 1) EER = HE ethanol HE ethanol+he gasoline 2) HE ethanol = m ethanol LHV 3) HE gasoline = m gasoline LHV Ethanol/Gasoline Energy Ratio is a method to analyze the relation between ethanol/gasoline ratios from the point of view of their energy. For instance, Huang, Hong [12] observe this relation to investigating the cooling effect of the in-cylinder pressure by altering the volume ratio of ethanol and gasoline. In fact, it is 5

14 concluded that an increase in the volume of ethanol in EER will decrease the incylinder temperature. Furthermore, EER was also considered to investigate the effect of the start of ignition on knock mitigation and effect of injection timing on lean combustion where two test conditions were studied [13]. The first one uses an EER equal to 24% and the other 48%. Ethanol/Gasoline Energy Ratio is important because this value permits to detect the best conditions to analyze engine s performance, such as energy efficiency and emissions. Similarly, Padala, Woo [14] have used an equation similar to Equation 1 to analyze an engine with ethanol and diesel injection. They define the ethanol fraction Ex in Equation 4. 4) Ethanol Fraction Ex = m ethanol CV ethanol m ethanol CV ethanol +m diesel CV diesel Where m represents the mass flow and CV the calorific value for ethanol and diesel. In that investigation, ethanol energy fraction is applied in the analysis of the indicative mean effective pressure, ignition delay and burn duration. Finally, it can be concluded that ethanol energy ratio is a parameter to review engine s performance in terms of the energy of the fuels involved in the test. It combines mass rates of two different fuels and their energy coefficients (for example low heating value or calorific value). 6

15 2.3 Mass Fraction Burn The mass fraction burn or MFB could be expressed as the rate of heat release in the combustion process [15]. One of the most used techniques to calculate MFB is one proposed by Rassweiler and Withrow [16] to obtain the apparent heat release. 5) dq apparent = γ γ 1 pdv + 1 γ 1 vdp In addition to the apparent heat release definition presented in Equation 5, there are other methods to determine the MFB. Yeliana, Cooney [15] acknowledge two methods: the single zone heat release and two zone model; nonetheless, it has been reported that the apparent heat release is widely accepted for its simplicity. Additionally, many tests have been performed in spark ignition engines to obtain the MFB curves for mixtures of gasoline and ethanol. Actually, Vinodh, Arvind [17] have studied and characterized MFB curves in a SI engine with single and split injection at 1500 rpm, partial load of 0.45 bar, 9.75 compression ratio and constant stoichiometric air-fuel ratio λ=1. Figure 3 shows MFB curves resulting from different fuels. For instance, fuel E85 and E50 demonstrated a slightly higher burn duration (27 and 28 crank angle degrees) than gasoline (26 crank angle degree) and faster flame expansion with the increase of ethanol content. Another study developed by Jiang, Ma [18] relates the speed of flame propagation and mass fraction burn for gasoline and alternative fuels. Indeed, the report shows the difference in flame propagation for ethanol and gasoline using dual fuel injection in an optical engine at a constant speed of 1200 rpm and 11.3 compression ratio. Besides, gasoline and ethanol were introduced separately 7

16 through port fuel injection and direct injection respectively. Figure 4 exhibits an upper flame speed curve for ethanol blend than gasoline that have led to conclude that ethanol induces a high flame propagation than gasoline that would contribute to enhancing efficiency. Similarly, Figure 5 illustrates the effect of ethanol and its contribution to flame growth compared with gasoline. Figure 3. MFB Curves for gasoline, E50, and E85. Single and split injection [19] Figure 4. Flame speed and MFB (IMEP 5 bar) [18] Likewise, Augoye and Aleiferis [20] have also studied MFB curves and flame development in a single cylinder optical engine at 1000 rpm and stoichiometric combustions. Several fuels like iso-octane, gasoline, E100, E96W6 (96% ethanol 8

17 and 6% water per volume), and E90W10 (90% ethanol and 10% water per volume) were employed using port fuel injection (PFI) and direct injection (DI). After injecting fuel through PFI, a crank angle degree (CAD) of 34.4 was observed at 50% MFB after ignition time (AIT) for E100. The same percentage of mass fraction burn was noticed for iso-octane at 40.8 CAD. Conversely, 37.6 and 42 CAD AIT at 50% MFB was observed for E100 and iso-octane respectively using DI [20]. Figure 6 shows mass fraction curves for the fuels involved in this research and contributes to conclude that anhydrous ethanol possess the best MFB curve in PFI and DI operations due to better flame growth in the combustion process. Figure 5. Flame images of Bio-fuels under dual-injection strategy (IMEP 5 bar) [18] 9

18 Figure 6. Mass fraction burned for PFI and DI operation [20] In conclusion, several investigations considering the influence of ethanol on mass fraction burn curves and flame propagation have demonstrated the benefit of this renewable fuel in spark ignition engines. Some of these investigations have been discussed in this section in which either ethanol or blends of this fuel has demonstrated similar or greater capabilities that regular fuels like gasoline or isooctane to improve combustion and engine performance. 2.4 Effect of Ethanol on Gas Emissions Ethanol has been preferred among some alcohols as one of the best options to replace common fuels used in engines [10]. High octane rating, oxygen content, relatively low carbon: hydrogen ratio and high latent heat are among its most respectable properties [21]. Various experiments have been performed by several researchers so that the most significant gas emissions like hydrocarbons (HC), carbon monoxide (CO), nitrogen oxide and nitrogen dioxide (also call NOx) can be examined [22]. Thus, it is significant to exhibit ethanol effects on SI engines gas emissions either using blends or pure ethanol. 10

19 Figure 7. Cumulative HC emission for blends E5 to E85 at 22 C [21] Firstly, hydrocarbon emissions can be reduced using ethanol thanks to a better combustion for the extra presence of oxygen in fuel chemical composition [21]. It has been observed a considerable decrease in gas emission for E5 to E50 fuels compare to unleaded gasoline as illustrated in Figure 7. Similarly, Koç, Sekmen [23] have investigated HC emissions in a single cylinder engine with two compression ratios (10:1 and 11:1) and speed increasing from 1500 to 5000 rpm. This research also attributes lower HC emissions due to ethanol characteristics like leaning effect or oxygen enrichment, Figure 8. Figure 8. Effect of unleaded gasoline-ethanol blends on HC emissions at different compression ratios [23] 11

20 Furthermore, Karavalakis, Short [24] have conducted a gas emission study in nine light vehicles (two of them fuel flexible vehicles) of different well-known brands and with distinct ethanol blends (E10, E15, E20, E51 & E83). They discovered that fuel E83 shows a 43 and 38 percent reduction in CO emissions compared with E51 and E10 respectively. They conclude that this effect is attributed to extra oxygen content in fuels with higher ethanol volumes. In other research, a single cylinder Honda engine 2.5 horsepower with a compression ratio of 6.1 was employed to assess gas emissions between ethanol and gasoline. The results show a 68 percent reduction of CO emissions [10]. Figure 9 presents a graph where we can appreciate CO emissions versus brake mean effective pressure. Repeatedly, there is a trend of ethanol to reduce CO emissions compared to gasoline. Conversely to HC and CO emissions, NOx emissions does not show a specific increase or decrease pattern [25]. For instance, Bielaczyc, Szczotka [21] comment that NOx emissions tend to rise due to the presence of oxygen in ethanol. Nevertheless, Li, Liu [26] have investigated a four stroke spark ignition engine using E100 at different air fuel ratios. They affirm multiple results like decrease in NOx emissions when delaying spark timing, reduction of NOx emissions in one of the richest air fuel mixtures (λ=0.85) and all the way round for a lean mixture (e.g. λ=1.05). Also, it has been observed that lower engine speeds result in lower NOx emissions compared to gasoline and the opposite when speed increases. The research concluded that at lower speeds ethanol heat of vaporization decreases cylinder temperature. This characteristic helps to decline NOx formation, but at high speed this feature exists but seems to have no considerable effect, Figure 10 and Figure

21 Figure 9. CO concentration versus BMEP at 20 BTDC [10] Similarly, Al-Farayedhi, Al-Dawood [27] investigated NOx formation in a sixcylinder engine at different engine loads and speed. They detected variations in NOx emissions for different fuels (E10, E15 & E20), speeds, and engine loads. To illustrate, ethanol blends in some cases reveal lower NOx emissions than unleaded gasoline ( base series in Figure 12), and higher NOx emission if engine load is changed as showed in Figure 13. Besides, they also discovered higher NOx emissions for ethanol blends than for unleaded gasoline keeping constant engine speed as presented in Figure 14. For those higher NOx emissions, they concluded that this is the effect of an increasing temperature in heating cycles with increased speed and residual gasses left in the combustion chamber. 13

22 Figure 10. Effects of lambda on NOx emissions versus spark timing [26] Figure 11. Comparison of NOx emissions for gasoline and ethanol [26] 14

23 Figure 12. NOx emissions vs. speed at a constant load of 340 kpa [27]. Figure 13. NOx emissions vs. speed at a constant load of 510 kpa [27]. 15

24 Figure 14. Effect of load on NOx emissions at a constant speed of 2000 rpm [27]. To summarize, ethanol has a marked impact on exhaust gas emissions in a spark ignition engine. Extra oxygen content in ethanol significantly reduces HC and CO emissions. On the other hand, oxygen content in ethanol has been linked with NOx formation at higher temperatures where the leaning effect of ethanol seems to be neglected. Despite variable results on NOx emission, ethanol fuel continues tempting scientists to obtain an environmentally friendly fuel for the present and future. 3. Experimental Procedure A former UTS Ph.D. conducted experiments in a modified engine with direct ethanol fuel injection and port fuel injection system. The pressure in the port fuel injection was 250 kpa and the one for direct ethanol fuel injection could be set between 3 and 13 MPa. An electronic control unit (ECU) managed the fuel systems already mentioned [11]. This data analyzed further in Section 4 was provided by the supervisor of this report 16

25 In addition, Figure 15 displays other equipment involved in the experiments like a dynamometer (number 2), pressure transducer (15), K-type thermocouples (8, 13 &18), buffer tank (19), lambda sensor (12) and gas analyzer (5). Moreover, Table 2 exhibits the characteristics of the engine used in this investigation. Figure 15. Schematic of the engine system [11]. Table 2. Specification of the engine [11]. Single cylinder Engine type air cooled 4 stroke Model Yamaha YBR 250 Cilinder Volume (cc) 249 Compression Ratio 9.8 Bore (mm) 74 Connecting rod length (mm) Crank radius (mm) 29 Ethanol delivery system Direct fuel injection Gasoline delivery system Port injection 17

26 4. Results and Discussion 4.1 Performance of SI Engine with Ethanol Direct Injection (EDI) and Gasoline Port Fuel Injection (PFI) p-v Diagrams Several pressure-volume diagrams have been plotted for different engine speeds using the information captured by the equipment and engine characteristics described in Section 3. Each speed considers some combinations for gasoline and ethanol fuel. Thus, four charts displaying pressure for y axis and volume on the x axis have been obtained. Figure 16 to Figure 19 contain p-v diagrams at different speeds as well as the distinct combination of ethanol and gasoline content. Identical colors have been used for diagrams with the same combination of gasoline and ethanol so that their evolution could be appreciated when analyzing different engine speeds. In Figure 16, we can appreciate a red curve that represents the p-v diagram using only gasoline as fuel. The diagrams vary when changing the fuel content. In fact, peak pressure rises with increasing ethanol content in the fuel at 3500 rpm engine speed. For instance, a dotted curve represents 8 milligrams of ethanol and 45 milligrams of gasoline. Besides, a black curve at the top of the diagrams possesses the highest ethanol content that is 13.4 milligrams. Figure 17 shows pressure-volume diagrams at 4000 revolutions per minute. In this graph, we are able to observe that a combination of gasoline 29 milligrams and 13.4 milligrams ethanol (black color) is no longer the one that reaches the highest pressure in the combustion stroke as it was at 3500 rpm. Instead, a 18

27 combination of 37 milligrams of gasoline and 10.7 milligrams of ethanol attain the highest pressure at this velocity. The red curve, which belongs to gasoline fuel, retains the lowest pressure in the combustion stroke. Figure 16. p-v diagrams: Different ethanol 3500 rpm Figure 17. p-v diagrams: Different ethanol 4000 rpm 19

28 Diagrams for pressure-volume at 4500 revolutions per minute are presented in Figure 18. In this plot, a combination of 45 and 8 milligrams of gasoline and ethanol respectively (dot curve) gets the highest pressure in the compression stroke. On the other hand, the maximum presence of ethanol in the mixed fuel (black curve) is momentarily located in between the curve that uses only gasoline and the one with a combination of 45 milligrams of gasoline and 8 milligrams of ethanol. Pressure-volume diagrams at this engine speed are closer than the ones at lower speeds. Figure 18. p-v diagrams: Different ethanol 4500 rpm 20

29 Figure 19. p-v diagrams: Different ethanol 5000 rpm Lastly, different curves are presented in Figure 19 at 5000 revolutions per minute. In this case, a combination of 49 milligrams of gasoline and 6.7 milligrams of ethanol gets the highest pressure at the compression stroke. Additionally, we could notice that the diagram for gasoline is not anymore the one with the lowest pressure. Instead, a mixture of 41 milligrams of gasoline and 9.4 milligrams of ethanol share this spot closely Indicated Mean Effective Pressure Indicated mean effective pressure for different speeds and fuel combinations have been obtained in this study. Also, Appendix 1 contains an evaluation to calculate the indicated mean effective pressure (IMEP) either by computing the average of IMEP or the average pressure. In this section, IMEP tends to increase with ethanol content for the majority of fuel combination as we could appreciate in Figure 20, Figure 22, Figure 24 and Figure

30 Figure 20. IMEP vs. ethanol 3500 rpm Figure 21. IMEP, Torque and Fuel content at 3500 rpm In Figure 20 and Figure 21, a fuel combination with 6.7 milligrams of ethanol is the only one that reaches a lower indicated mean effective pressure than its 22

31 predecessor. The speed in which this behavior was observed is 3500 rpm. Figure 22 and Figure 23, on the other hand, illustrate more variety at 4000 rpm. For instance, 5.5 and 6.7 milligrams of ethanol in the fuel combination got a lower IMEP than that with 4 milligrams. Higher and lower IMEP are appreciated at distinct gasoline and ethanol content. The highest value obtained for this set of data results with 10.7 milligrams of ethanol. Figure 22. IMEP vs. ethanol 4000 rpm 23

32 Figure 23. IMEP and torque at 4000 rpm At 4500 rpm, a trend to increase IMEP with ethanol content can be seen in Figure 24 and Figure 25. Eight milligrams of ethanol in the fuel combination increases the IMEP roughly 2.5 percent the pressure of the previous mix. This fuel combination is the second highest at this speed. The greatest IMEP occurred at 13.4 milligrams of ethanol in the fuel combination. Lastly, a speed of 5000 RPM gave the date showed in Figure 26 and Figure 27. In these figures, there is a continuous tendency to augment IMEP with increase of ethanol content; nonetheless, a combination of 5.5 milligrams of ethanol in the fuel combination dropped its pressure to a lower value than the previous one. Similar performance can be observed for data after 6.7 milligrams of ethanol. For this data, the highest value happened with a combination containing 13.4 milligrams of ethanol. 24

33 Figure 24. IMEP vs rpm Figure 25. IMEP and torque at 4500 rpm 25

34 Figure 26. IMEP and torque at 4500 rpm. Figure 27. IMEP vs. ethanol 5000 rpm. Overall, one might say that higher IMEP values have been observed for fuel mixtures fuels that contain ethanol. Among the multiple combination of ethanol 26

35 and gasoline fuels, some curves may present an undulating behavior as illustrated in Figure 28. Nonetheless, any fuel containing ethanol in the data analyzed have reached a higher IMEP than gasoline itself. Thus, this fact supports a previous section in this report where this characteristic was acknowledged when ethanol fuel is employed in spark ignition engines. Figure 28. Summary IMEP vs. Ethanol content at different RPM 27

36 4.1.3 Brake Mean Effective Pressure Figure 29. Variation of BMEP with EER [11]. In Figure 29, brake mean effective pressure (BMEP) curves have been plotted by Zhuang and Hong [11] research. The set of curves in the above figure are grouped into two categories, such as light and medium load. This section would compare and analyze medium load data. Medium load data have been zoomed in on Figure 30 to observe them easier. It can be noticed that an increase in ethanol energy ratio (EER) also augment BMEP. Values from 0.60 MPa to 0.68 MPa can be found in medium load curves for this figure. However, using base data and Equation 6 results in 12.5 % higher BMEP outcomes as showed in Figure 31. Figure 30 and Figure 31 trajectories are exactly the same except that Figure 31 presents larger BMEP values at each EER. Despite the difference in results, characteristics like a high latent heat of vaporization, high combustion velocity, and mole multiplier effect have been considered to rise BMEP [11]. 6) BMEP = Pb n V d N 28

37 Figure 30. Zoom in on Figure 29 - Medium load. Figure 31. BMEP vs. EER Medium load using Equation Volumetric Efficiency Volumetric efficiency data is illustrated in Figure 32. Using base data spreadsheet from the research in analysis has allowed to compute and draw Figure 33. In this figure, trajectories for every speed coincide except one point located on the curve for 3500 rpm at 48.4 % EER. 7) n v = n m a ρ a V d N 29

38 Figure 32. Variation of Volumetric efficiency with EER [11]. Figure 33. Zoom in on Figure 32 - Medium load. Equation 7 is applied to confirm volumetric efficiency results and visualize them in a graph. Therefore, Figure 34 points out a remarkable difference in trajectories compared with the ones in Figure 33. Nevertheless, the speed at which the base data was taken is not precisely 3500, 4000, 4500 and 5000 rpm, so the real speed was put to use in order to be more consistent with computing the results. Hence, Figure 35 was achieved with two important characteristics. The first one is that the trajectory of points matches with the ones in Figure 33, and the other one is that the volumetric efficiency of every single point is 5.6% higher. Zhuang and Hong [11] documented volumetric efficiencies at different speeds varying from 67 30

39 to 72 %; nonetheless, volumetric efficiencies ranging from 71 to 77 % can be read in Figure 35. Figure 34. Volumetric efficiency vs. EER - Medium load using Equation 7. Figure 35. Volumetric efficiency vs. EER - Medium using Equation 7 and real RPM Brake Specific Fuel Consumption Brake specific fuel consumption (BSFC) trajectories for medium and light load are illustrated in Figure 36. An enlargement of medium load points is exhibited in Figure 37. On the other hand, Figure 38 uses Equation 8 to determine BSFC curves. In this figure, it is possible to see that the trajectory of the curves differs 31

40 from the ones in Zhuang and Hong [11] research. In order to solve this inconvenient, real speeds were applied in Equation 8. The results can be inspected in Figure 39. The lines in this figure agree exactly with the original ones. There is no variation of any value at each point. 8) bsfc = mf Pb Figure 36. Variation of BSFC with EER [11]. Figure 37. Zoom in on Figure 36 - Medium load. 32

41 Figure 38. BSFC vs. EER - Medium load using Equation 8. Figure 39. BSFC vs. EER Medium load using Equation 8 and real RPM Brake Specific Emissions Brake specific gas emission for CO, NO and HC were studied by Zhuang and Hong [11] using Ethanol Direct Injection (EDI) and Gasoline Port Injection (GPI). In this section, brake specific gas emission have been plotted utilizing medium load base data. Furthermore, an analysis of the results obtained in Zhuang s work and comparisons with other studies have been made. 33

42 Figure 40. Variation of BSCO with EER [11]. Figure 41. BSCO vs. EER Medium load. To start, Figure 40 illustrates the curves obtained for light and medium load at different engine speeds in the study above mentioned. Besides, Figure 41 shows only brake specific CO emissions (BSCO) for medium load data. In the graphs, there is a trend to increase carbon monoxide gas emissions with an increase in the content of ethanol. The greater the content of ethanol (or EER) in the fuel the higher the amount of CO emissions. Similarly, there is an increase in the content of gas emission for hydrocarbons (HC) as we can appreciate in Figure 42 and Figure 43. This pattern has been imputed to a difficulty for the flame to propagate 34

43 with the increase of Ethanol; in fact, Huang, Hong [12] found an overcooled region close to the cylinder wall that provoke an increase in CO and HC emissions. Likewise, Brewster, Railton [29] acknowledge an increase in HC emissions in a study using EDI. Despite of several factors that can contribute to increasing HC emissions such as, flame quenching, crevice filling, absorption-desorption of oil layer and incomplete combustion, Brewster, Railton [29] argue that direct injection complicate the scenario conferring a certain lack of homogeneity and soaking of chamber surfaces. Figure 42. Variation of BSHC with EER [11]. Figure 43. BSHC vs. EER Medium load. 35

44 Conversely to BSCO and BSHC, BSNO gas emissions decrease with an increase of EER as appreciated in Figure 44 and Figure 45. This behavior has also been seen in other research that has attributed it to a diminish of temperature in the combustion process thanks to the use of ethanol [30]. Figure 44. Variation of BSNO with EER [11]. Figure 45. BSNO vs. EER Medium load. Finally, no trajectory differences between the graphs found in the research in analysis and figures for brake specific emissions drawn in this section have been observed. 36

45 4.2 Ethanol Effect on Burning duration, Temperature Drop, and Compression Ratio Mass Fraction Burn and Ethanol Energy Ratio The burning process varies from 40 to 60 crank angle degrees (CAD) depending on the engine design and operation [31]. The start and end of combustion can be determined using a logarithmic diagram pressure versus volume. Figure 46, for instance, clearly reveals the polytrophic compression and expansion slopes [32]. Likewise, pressure one and three which are the continuation of compression and expansion lines are identified in this graph [33]. This information has been used in combination with Equation 9 and 10 to plot MFB curves. 9) P 2 = P θ ( V V c ) n 10) x b = P 2 P 1 P 3 P 2 where: P 2 : P θ : V: V c : n: Projection of instantaneous pressure on minimum volume line Instantaneous pressure Instantaneous volume Clearance Volume Polytropic index 37

46 Figure 46. Log p vs. Log v; 3500 rpm; G29 E13.4 Figure 47 exhibits MFB curves at 3500 rpm for distinct combinations of ethanol and gasoline. In this figure, gasoline fuel is the lowest curve on the chart while G57E4 is the highest line in this band of curves. It is fundamental to observe that the minimum presence of ethanol amongst fuel combinations, in this case, G57E4, gets the fastest mass burn trend; in fact, this MFB line is even found higher than those containing more ethanol in the fuels tested. Furthermore, Table 3 indicates combustion time delays and combustion time periods. A graphical representation of this information could be seen in Figure 48 and Figure 49. First, the lowest combustion time delay is observed at 19 % EER in Figure 48. Higher EER than the one mention previously increases the crank angle variation. Conversely, combustion time periods represented in Figure 49 achieve lower crank angle variations with any increase in ethanol energy ratio. In similar tendency, the lowest combustion time period has been reached with 19% 38

47 EER. Thus, an increase in ethanol content enhances combustion and contributes to burning fuel in fewer crank angles in comparison with gasoline fuel. G57 E4 G69 E0 Figure 47. MFB curves at 3500 rpm. Table 3. Combustion time delay and period at 3500 rpm. Crank Angle Variation N Sample EER Combustion Time Delay Combustion Time Period RPM Code % Δ (0-5%) Δ (0-10%) Δ (10-90%) Δ (5-90%) 3500 G69 E G57 E G52.5E G49 E G45 E G41 E G37 E G29 E

48 Figure 48. Crank angle variation vs. EER at 3500 rpm: Time delay Figure 49. Crank angle variation vs. EER at 3500 rpm: Combustion period Mass fraction burn trajectories at 4000 rpm can be seen in Figure 50. In these set of curves, the lowest limit of the band of lines occurs at G29E13.4 while the highest limit is shared between G57E4 and G41E9.4. Other MFB curves can be found between the lowest and topmost lines. Conversely to the curves at 3500 rpm, these curves are crossing each other at a point located between 30 to 50% of mass fraction burn. Additionally, Table 4 summarizes combustion time delays and combustion time periods at this speed. Figure 51 and Figure 52 also present this information graphically. Figure 51, for example, indicates that values of 19 and 26 % EER 40

49 achieved lower time delays compared with other energies. When compared with gasoline fuel, a 19% EER gets 10.7% lower CAD variation for Δ (0-5%) line; similarly, 26% EER get 11.7% lower CAD for Δ (0-10%) line. Likewise, Figure 52 displays combustion time periods. The minimum value at this speed was observed at 43% EER with 25.9 crank angle degrees for Δ (10-90%) and 28.5 crank angle degrees for Δ (5-90%). Repeatedly, it is observed that any energy containing ethanol make smaller crank angle intervals than that of gasoline fuel. G41 E9.4 G57 E4 G29 E13.4 Figure 50. MFB curves at 4000 rpm. Table 4. Combustion time delay and period at 4000 rpm. Crank Angle Variation N Sample EER Combustion Time Delay Combustion Time Period RPM Code % Δ (0-5%) Δ (0-10%) Δ (10-90%) Δ (5-90%) 4000 G69 E G57 E G52.5E G49 E G45 E G41 E G37 E G29 E

50 Figure 51. Crank angle variation vs. EER at 4000 rpm: Time delay Figure 52. Crank angle variation vs. EER at 4000 rpm: Combustion period G29 E13.4 G57 E4 Figure 53. MFB curves at 4500 rpm. 42

51 The set of MFB curves in Figure 53 correspond to 4500 rpm engine speed. Two fuels containing ethanol limit the group of traces illustrated. These curves are G29E13.4 highest limit and G57E4 lowest limit. Likewise, significant information regarding these curves has been obtained in Table 5 as well as graphically in Figure 54 and Figure 55. Table 5. Combustion time delay and period at 4500 rpm. Crank Angle Variation N Sample EER Combustion Time Delay Combustion Time Period RPM Code % Δ (0-5%) Δ (0-10%) Δ (10-90%) Δ (5-90%) 4500 G69 E G57 E G52.5E G49 E G45 E G41 E G37 E G29 E Figure 54. Crank angle variation vs. EER at 4500 rpm: Time delay. In Figure 54, we might separate CAD variations in two segments. The first one would group 19 and 26 % EER. These two energies get the highest CAD variation in Δ (0-5%) and Δ (0-10%). Crank angle intervals for these energies are up to 16 percent higher at Δ (0-10%) and 20 percent higher at Δ (0-5%) than gasoline fuel itself. The other segment would involve energies between 31 to 61 % EER. Conversely, an increase in EER in this segment decreases CAD variations in 43

52 comparison with gasoline fuel. This portion possesses the lowest CAD variation that are 6 CAD for the Δ (0-5%) line and 9 CAD at 49% EER for the Δ (0-10%) line. Combustion time periods at 4500 rpm exhibit variation along Δ (10-90%) and Δ (5-90%) lines. These representations are on view in Figure 55. The minimum CAD variation is 36.2 in line Δ (10-90%). This crank angle variation occurs at 43 and 61 EER. Line Δ (5-90%), on the other hand, reaches its lowest value at 43 EER with 39.8 CAD variation. Finally, this figure does not show a specific pattern with an increase in EER. For example, one could observe high and low crank angle values independently of the quantity of ethanol in the fuel. Figure 55. Crank angle variation vs. EER at 4500 rpm: Combustion period. Lastly, MFB curves are plotted in Figure 56 which corresponds to 5000 rpm engine speed. We would say that the band of curves looks quite organized in this figure in comparison with the MFB curves at 4000 and 4500 rpm. Moreover, the lines corresponding to G69E0 and G57E4 are overlapped on each other in the lowest limit of these curves. Fuel G41E9.4 rules the highest limit. Also, this figure clearly shows that an increase in ethanol content is enhancing combustion and moving MFB curves up from the one equivalent to gasoline fuel. 44

53 G41 E9.4 G69 E0 & G57 E4 Figure 56. MFB curves at 5000 rpm. Table 6. Combustion time delay and period at 5000 rpm. Crank Angle Variation N Sample EER Combustion Time Delay Combustion Time Period RPM Code % Δ (0-5%) Δ (0-10%) Δ (10-90%) Δ (5-90%) 5000 G69 E G57 E G52.5E G49 E G45 E G41 E G37 E G29 E Figure 57. Crank angle variation vs. EER at 5000 rpm: Time delay. 45

54 Figure 57 shows a tendency to decrease CAD variation with an increase in EER. This peculiarity occurs in most of the points except at 37% EER in the Δ (0-5%) line and 26% EER in the Δ (0-10%) line. Besides, the lowest CAD variation is six crank angle at 49% EER in the Δ (0-5%) line and 9.8 crank angle at 43% EER in the Δ (0-10%) line. On the other hand, combustion time periods are exhibited in Figure 58 where the lowest crank angle variations take place at 31% EER in Δ (10-90%) and Δ (5-90%) lines. No obvious tendency can be perceived in these two variation lines; nevertheless, we might say that some fuels with ethanol content have reached lower CAD variation in comparison with gasoline. Figure 58. Crank angle variation vs. EER at 5000 rpm: Combustion period. Figure 59. Mass fraction burned for PFI and DI operation [20]. 46

55 To conclude, most of the observations in this section agree with the facts investigated in section 2.3. For instance, the MFB patterns revealed at 3500 and 5000 rpm (Figure 47 & Figure 56) are quite similar to the ones obtained in a flame development research using several fuels including gasoline and ethanol port fuel and direct injection. The MFB curves of this research (Figure 59) show gasoline fuel standing in the lowest limit of the graph while fuels containing ethanol reach upper spots thanks to a better flame growth in the combustion process [20]. Accordingly, other engine speeds in our data analyzed, such as 4000 and 4500 have also demonstrated improved MFB characteristics in the majority of EER tested as noticed in Figure 52 and Figure 55 for 4000 and 4500 rpm respectively. Otherwise, some tested fuels that increased their burning duration in comparison with gasoline were noticed in Figure 55 (4500 rpm) and Figure 58 (5000 rpm). Similar results were also observed in a research where fuels containing ethanol slightly increase their burn duration when compared with gasoline [17] Ignition Time Delay The effect of ethanol on the ignition time delay has also been investigated in this section. It is fundamental to develop this analysis and compare CAD at different engine speeds. These would provide a more clear understanding of ethanol effect on ignition delay. Therefore, Figure 60 and Figure 61 have been plotted to analyze these changes at different EER. Firstly, the lines observed in Figure 60 represent CAD obtained from 0-5% MFB data intervals. This set of curves showed an interesting characteristic. For instance, lines that reached higher CADs at 3500 and 4000 rpm conversely drop to lower CADs at higher engine speeds such as 4500 and 5000 rpm. For example, 49% EER gets 7.7 CAD at 4000 rpm but 6.2 CAD at 4500 rpm. In the same trend, 47

56 this behavior could be observed for several ethanol energy ratios. Nevertheless, it is believed that a better analysis could be done by examining time delays at every single engine speed. Figure 60. CAD (0-5%MFB) vs. RPM One of the speeds in Figure 60 is 3500 rpm. In this case, the majority of EERs stand in a higher point when compared with EER 0%. The only exception at this speed occurs on EER 18%. This point reveals 4.2 CAD variation that is lower than 5.2 CAD for gasoline. Likewise, two points are below the one for gasoline fuel at 4000 rpm. EER 18 and 25% have shown 5 CAD at this speed that is 10.7% lower than gasoline fuel. At 4500 rpm, EER varying from 31 to 61 have fallen under the point for gasoline fuel. The lowest CAD illustrated is six. This CAD is share for EER 31 and 61 percent. The decrease in time delay compared with gasoline is 11.7%. Finally, ethanol energy ratio ranging from 25 to 61% are below EER 0%. Only EER 18% is higher than that for gasoline. The lowest CAD seen is 6.1 which belongs to EER 49%. A 21.7% drop compared with gasoline reveals that this point has attained the highest fall of all the engine speeds involved in Figure

57 Figure 61. CAD (0-10%MFB) vs. RPM On the other hand, Figure 61 presents information regarding CAD from 0-10% MFB data intervals. Figure 61 resembles Figure 60 with only a few differences. First, EER 31 has dropped below EER 0 and has joined EER 18 that was the only observed before in Figure 60. The minimum CAD at this speed is still seized by EER 18. This spot is 18.8% lower than gasoline. At 4000 rpm, there are no significant changes as remarked in Figure 60. The lowest location at this speed is shared again for EER 18 and 25. They represent an 11.7 percent drop compared with gasoline. Similarly, no changes are seen when engine speed reached 4500 rpm. The minimum CAD represents 14.2% drop at this speed. Lastly, there is no point above the one for gasoline at 5000 rpm. The lowest drop represents a 16.9 % fall related to gasoline. Based on the above observations and analysis, we would say that two scenarios could be presented in this section. The first one will be that lower speed reached inferior CAD variations for 0-5% and 0-10% MFB. To illustrate, crank angle degrees, such as 4.2, 5, 6 and 6.10 were plotted at 3500, 4000, 4500 and 5000 rpm respectively (Figure 60). The other scenario indicates that time delay seems to be affected when a higher presence of ethanol is injected in the fuel at lower speeds. 49

58 For example, most of the EER are above the line for gasoline fuel at 3500 and 4000 rpm (Figure 60 and Figure 61) while few of them are above gasoline fuel at 4500 and 5000 rpm Combustion Time Period In order to analyze the combustion time period at different EER, Figure 62 has been depicted so that CAD versus engine speed could be examined. First of all, a considerable gap between gasoline fuel and other fuels containing ethanol could be seen in Figure 62 at 3500 rpm. EER varying from 18 to 61 are group very closely in a sector between 27.9 and 29 CAD. The lowest value indicates 17.7% drop compared with gasoline. This lowest value belongs to EER 18. Figure 62. CAD (10-90%MFB) vs. RPM Furthermore, there is no fuel with ethanol that overpass gasoline crank angle at 4000 rpm. In contrast to the trend observed at 3500 rpm, EER curves are spread in CAD values from 25.9 to The lowest CAD is 25.9 for EEE 43 at 4000 rpm. A 15.9% drop is registered in Figure 62 when compared to gasoline. 50

59 Additionally, EER 18 and 49 have shown to gain higher locations than the one occupied by gasoline fuel at 4500 rpm; on the contrary, EER 25, 31, 37, 43 and 61 are lower than 39.7 CAD. The lowest value among these energies is 12.6% inferior to gasoline fuel. Finally, all the fuels that contain ethanol except EER 18 have accomplished lower crank angle variation than gasoline fuel at 5000 rpm. The smallest value of the energies is been detected at 26.5 CAD. This value is 34.5 % lesser than the one for EER 0. To sum up, crank angle degrees for fuels containing ethanol have a general tendency to reduce the time in which the combustion time period is completed. For instance, an ethanol content of 9.4 milligrams of ethanol in the fuel has fulfilled the 10-90% MFB in only 25.9 CAD. Another example is the fuel consisting of 4 milligrams of ethanol in the fuel (G57 E4) completed the same interval of MFB in 27.9 CAD. To conclude, the combustion time period has been improved in most of the cases. This betterment could be attributed to ethanol s characteristic to enrich flame growth as reported in section 2.3 of this document Temperature Drop at the End of Combustion Process. The theoretical temperature at the start and end of the combustion (T2 & T3) is showed by the diagram temperature- entropy in Figure 63. Using the data analyzed in this section, T2 & T3 have been calculated for every combination of ethanol and gasoline fuel so that the temperature drop at the end of the combustion process could be determined. The theoretical temperature at the start of the combustion process could be calculated using Equation

60 Figure 63. Representation of the constant-volume cycle in the p, v and T, s diagram [34]. 11) T 2 = (r v ) k 1 where: r v : k: Compression Ratio Specific Heat Ratio In order to calculate T2, the specific heat ratios for every single combination of ethanol and gasoline were determine from the diagrams log pressure versus log volume used in section Besides, the measured temperature at the start of compression is 21.6 Celsius degrees and the compression ratio 9.8. Therefore, Table 7, Table 8, Table 9 and Table 10 present the theoretical temperature at the start of the combustion at 3500, 4000, 4500 and 5000 rpm respectively. 52

61 Table 7. Theoretical temperature at start of combustion: 3500 rpm N Sample EER k compression T2 RPM Code % K 3500 G29 E G37 E G41 E G45 E G49 E G52.5 E G57 E G69 E Table 8. Theoretical temperature at start of combustion: 4000 rpm N Sample EER k compression T2 RPM Code % K G29 E G37 E G41 E G45 E G49 E G52.5 E G57 E G69 E Table 9. Theoretical temperature at start of combustion: 4500 rpm N Sample EER k compression T2 RPM Code % K G29 E G37 E G41 E G45 E G49 E G52.5 E G57 E G69 E Table 10. Theoretical temperature at start of combustion: 5000 rpm N Sample EER k compression T2 RPM Code % K G29 E G37 E G41 E G45 E G49 E G52.5 E G57 E G69 E

62 The theoretical maximum temperature in the combustion process might be obtained by Equation ) T 3 = HV C V ( A + T F +1) 2 where: HV: c v : Heating value [MJ/kg] Specific Heat at Constant Volume [J/kg-K] A F : Air Fuel Ratio Since the majority of fuels in this analysis are a combination of gasoline and ethanol, the heating value of the mixed fuel is calculated with Equation ) HV fuel = m ethanol HV ethanol +m gasoline HV gasoline m ethanol +m gasoline where: HV ethanol : HV gasoline : Heating Value Ethanol = 26.9 [MJ/kg] Heating Value Gasoline = 42.9 [MJ/kg] m ethanol : Ethanol Mass Flow [kg/h] m gasoline : Gasoline Mass Flow [kg/h] 54

63 Likewise, the specific heat at constant volume needs to be obtained for the distinct combinations of fuels. Equation 14, 15 and 16 are employed to determine this value required towards obtaining the theoretical maximum temperature in the combustion process. Hence, calculations of the theoretical maximum temperature (T3) are displayed in Table 11, Table 12, Table 13 and Table 14 for different engine speeds. 14) c v fuel = m ethanol c v ethanol +m gasoline c v gasoline m ethanol +m gasoline 15) c v gasoline = c p gasoline k 16) c v ethanol = c p ethanol k where: c p gasoline : c p ethanol : k: Gasoline Specific Heat at Constant Pressure = 2041 [J/kg-K] Ethanol Specific Heat at Constant Pressure = 2339 [J/kg-K] Specific Heat Ratio Compression slope in log P vs log V diagram m gasoline : Gasoline Mass Flow [kg/h] m ethanol : Ethanol Mass Flow [kg/h] 55

64 Table 11. Temperature at End of Combustion 3500 rpm Sample EER Fuel flow G Fuel Flow E Total HV Fuel Cv fuel AFR T3 Code % Kg/h Kg/h Kg/h MJ/kg kj/kg-k K 1 G29 E G37 E G41 E G45 E G49 E G52.5 E G57 E G69 E Table 12. Temperature at End of Combustion 4000 rpm Sample EER Fuel flow G Fuel Flow E Total HV Fuel Cv fuel AFR T3 Code % Kg/h Kg/h Kg/h MJ/kg kj/kg-k K 1 G29 E G37 E G41 E G45 E G49 E G52.5 E G57 E G69 E Table 13. Temperature at End of Combustion 4500 rpm Sample EER Fuel flow G Fuel Flow E Total HV Fuel Cv fuel AFR T3 Code % Kg/h Kg/h Kg/h MJ/kg kj/kg-k K 1 G29 E G37 E G41 E G45 E G49 E G52.5 E G57 E G69 E Table 14. Temperature at End of Combustion 5000 rpm Sample EER Fuel flow G Fuel Flow E Total HV Fuel Cv fuel AFR T3 Code % Kg/h Kg/h Kg/h MJ/kg kj/kg-k K 1 G29 E G37 E G41 E G45 E G49 E G52.5 E G57 E G69 E

65 The tables displayed above are more clearly expressed in Figure 64, Figure 65, Figure 66 and Figure 67. In these figures, one could observe that the maximum temperature is reached by gasoline fuel (EER 0). Also, it is possible to notice that an inclusion of ethanol fuel decreases the temperature at the end of the combustion. Indeed, the higher concentration of ethanol in the fuel, the lower the temperature drop. The maximum temperature drops occur with EER 61 at 3500, 4000, 4500 and 5000 rpm. Figure 64. Temperature Drop 3500 rpm Figure 65. Temperature Drop 4000 rpm 57

66 Figure 66. Temperature Drop 4500 rpm Figure 67. Temperature Drop 5000 rpm In summary, the use of ethanol in the fuel contributes to decreasing the temperature at the end of the combustion and benefits to avoid the knock effect. The drops have got up to 25.3 %, 30.5%, 24.5% and 19.1% temperature reduction at 3500, 4000, 4500 and 5000 rpm respectively. Thus, the next steps consist of estimating what would be the equivalent compression ratio for fuels containing ethanol once the temperature drops have been already computed. 58

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