Ethanol Fuel as Renewal Energy in Spark
Ignition Engines
Julio César Villavicencio Cevallos
Master of Engineering
Manufacturing Engineering and
Management
2015
Table of contents
Table of contents ............................................................................................... i
List of Figures .................................................................................................. ii
List of Tables .................................................................................................... v
Appendix .......................................................................................................... v
Abstract ........................................................................................................... vi
1.
Introduction ................................................................................................. 1
2.
Literature Review ....................................................................................... 2
2.1 Specific Heat Ratio and Compression Ratio in Engine Thermal Efficiency
using Ethanol ................................................................................................... 2
2.2 Ethanol/ Gasoline Energy Radio ............................................................... 5
2.3 Mass Fraction Burn ................................................................................... 7
2.4 Effect of Ethanol on Gas Emissions ........................................................ 10
3.
Experimental Procedure ............................................................................ 16
4.
Results and Discussion ............................................................................. 18
4.1 Performance of SI Engine with Ethanol Direct Injection (EDI) and
Gasoline Port Injection (PFI) ........................................................................ 18
4.1.1 p-V Diagrams .................................................................................... 18
4.1.2 Indicated Mean Effective Pressure ................................................... 21
4.1.3 Brake Mean Effective Pressure ........................................................ 28
4.1.4 Volumetric Efficiency ...................................................................... 29
4.1.5 Brake Specific Fuel Consumption .................................................... 31
4.1.6 Brake Specific Emissions ................................................................. 33
4.2 Ethanol Effect on Burning duration, Temperature Drop, and Compression
Ratio .............................................................................................................. 37
i
4.2.1 Mass Fraction Burn and Ethanol Energy Ratio ................................ 37
4.2.2 Ignition Time Delay ......................................................................... 47
4.2.3 Combustion Time Period.................................................................. 50
4.2.4 Temperature Drop at the End of Combustion Process. .................... 51
4.2.5 Estimation of Compression Ratio..................................................... 59
5.
Conclusions............................................................................................... 62
6.
Appendix ................................................................................................... 64
7.
References & Bibliography ...................................................................... 68
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] ...................... 10
Figure 7. Cumulative HC emission for blends E5 to E85 at 22 °C [21] ........ 11
Figure 8. Effect of unleaded gasoline-ethanol blends on HC emissions at
different compression ratios [23] .................................................................... 11
Figure 9. CO concentration versus BMEP at 20° BTDC [10] ....................... 13
Figure 10. Effects of lambda on NOx emissions versus spark timing [26].... 14
Figure 11. Comparison of NOx emissions for gasoline and ethanol [26] ...... 14
Figure 12. NOx emissions vs. speed at a constant load of 340 kPa [27]. ...... 15
Figure 13. NOx emissions vs. speed at a constant load of 510 kPa [27]. ...... 15
ii
Figure 14. Effect of load on NOx emissions at a constant speed of 2000 rpm
[27]. ................................................................................................................. 16
Figure 15. Schematic of the engine system [11]. ........................................... 17
Figure 16. p-V diagrams: Different ethanol content @ 3500 rpm ................. 19
Figure 17. p-V diagrams: Different ethanol content @ 4000 rpm .................. 19
Figure 18. p-V diagrams: Different ethanol content @ 4500 rpm .................. 20
Figure 19. p-V diagrams: Different ethanol content @ 5000 rpm ................. 21
Figure 20. IMEP vs. ethanol content @ 3500 rpm ......................................... 22
Figure 21. IMEP, Torque and Fuel content at 3500 rpm ............................... 22
Figure 22. IMEP vs. ethanol content @ 4000 rpm ......................................... 23
Figure 23. IMEP and torque at 4000 rpm ....................................................... 24
Figure 24. IMEP vs. ethanol @ 4500 rpm ..................................................... 25
Figure 25. IMEP and torque at 4500 rpm ....................................................... 25
Figure 26. IMEP and torque at 4500 rpm. ...................................................... 26
Figure 27. IMEP vs. ethanol content @ 5000 rpm. ........................................ 26
Figure 28. Summary IMEP vs. Ethanol content at different RPM................. 27
Figure 29. Variation of BMEP with EER [11]. .............................................. 28
Figure 30. Zoom in on Figure 29 - Medium load. .......................................... 29
Figure 31. BMEP vs. EER – Medium load using Equation 6. ....................... 29
Figure 32. Variation of Volumetric efficiency with EER [11]....................... 30
Figure 33. Zoom in on Figure 32 - Medium load. .......................................... 30
Figure 34. Volumetric efficiency vs. EER - Medium load using Equation 7. 31
Figure 35. Volumetric efficiency vs. EER - Medium using Equation 7 and real
RPM. ............................................................................................................... 31
Figure 36. Variation of BSFC with EER [11]. ............................................... 32
Figure 37. Zoom in on Figure 36 - Medium load. .......................................... 32
Figure 38. BSFC vs. EER - Medium load using Equation 8. ......................... 33
Figure 39. BSFC vs. EER – Medium load using Equation 8 and real RPM. . 33
Figure 40. Variation of BSCO with EER [11]. .............................................. 34
Figure 41. BSCO vs. EER Medium load........................................................ 34
Figure 42. Variation of BSHC with EER [11]. .............................................. 35
iii
Figure 43. BSHC vs. EER Medium load........................................................ 35
Figure 44. Variation of BSNO with EER [11]. .............................................. 36
Figure 45. BSNO vs. EER Medium load. ...................................................... 36
Figure 46. Log p vs. Log v; 3500 rpm; G29 E13.4 ........................................ 38
Figure 47. MFB curves at 3500 rpm. ............................................................. 39
Figure 48. Crank angle variation vs. EER at 3500 rpm: Time delay ............. 40
Figure 49. Crank angle variation vs. EER at 3500 rpm: Combustion period. 40
Figure 50. MFB curves at 4000 rpm. ............................................................. 41
Figure 51. Crank angle variation vs. EER at 4000 rpm: Time delay ............. 42
Figure 52. Crank angle variation vs. EER at 4000 rpm: Combustion period. 42
Figure 53. MFB curves at 4500 rpm. ............................................................. 42
Figure 54. Crank angle variation vs. EER at 4500 rpm: Time delay. ............ 43
Figure 55. Crank angle variation vs. EER at 4500 rpm: Combustion period. 44
Figure 56. MFB curves at 5000 rpm. ............................................................. 45
Figure 57. Crank angle variation vs. EER at 5000 rpm: Time delay. ............ 45
Figure 58. Crank angle variation vs. EER at 5000 rpm: Combustion period. 46
Figure 59. Mass fraction burned for PFI and DI operation [20]. ................... 46
Figure 60. CAD (0-5%MFB) vs. RPM ......................................................... 48
Figure 61. CAD (0-10%MFB) vs. RPM ........................................................ 49
Figure 62. CAD (10-90%MFB) vs. RPM ...................................................... 50
Figure 63. Representation of the constant-volume cycle in the p, v and T, s
diagram [34]. ................................................................................................... 52
Figure 64. Temperature Drop – 3500 rpm .................................................... 57
Figure 65. Temperature Drop – 4000 rpm ..................................................... 57
Figure 66. Temperature Drop – 4500 rpm ..................................................... 58
Figure 67. Temperature Drop – 5000 rpm ..................................................... 58
iv
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]. ...................................................... 17
Table 3. Combustion time delay and period at 3500 rpm. ............................. 39
Table 4. Combustion time delay and period at 4000 rpm. ............................. 41
Table 5. Combustion time delay and period at 4500 rpm. ............................. 43
Table 6. Combustion time delay and period at 5000 rpm. ............................. 45
Table 7. Theoretical temperature at start of combustion: 3500 rpm .............. 53
Table 8. Theoretical temperature at start of combustion: 4000 rpm .............. 53
Table 9. Theoretical temperature at start of combustion: 4500 rpm .............. 53
Table 10. Theoretical temperature at start of combustion: 5000 rpm ............ 53
Table 11. Temperature at End of Combustion – 3500 rpm ............................ 56
Table 12. Temperature at End of Combustion – 4000 rpm ............................ 56
Table 13. Temperature at End of Combustion – 4500 rpm ............................ 56
Table 14. Temperature at End of Combustion – 5000 rpm ............................ 56
Table 15. Estimated Compression Ratio: 3500 rpm....................................... 60
Table 16. Estimated Compression Ratio: 4000 rpm....................................... 60
Table 17. Estimated Compression Ratio: 4500 rpm....................................... 60
Table 18. Estimated Compression Ratio: 5000 rpm....................................... 60
Appendix
Appendix 1. Calculation of IMEP using average pressure. ........................... 64
v
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
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
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
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 12.78 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
9.2
36.5
39.3
11.85
36.4
40.9
12.78
34.8
42.6
2
Δ
2.8
4.5
7.8
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
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
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.
̇ 𝑒𝑡ℎ𝑎𝑛𝑜𝑙
𝐻𝐸
̇ 𝑒𝑡ℎ𝑎𝑛𝑜𝑙 +𝐻𝐸
̇ 𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒
𝐻𝐸
1)
𝐸𝐸𝑅 =
2)
̇ 𝑒𝑡ℎ𝑎𝑛𝑜𝑙 = 𝑚̇𝑒𝑡ℎ𝑎𝑛𝑜𝑙 ∙ 𝐿𝐻𝑉
𝐻𝐸
3)
̇ 𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒 = 𝑚̇𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒 ∙ 𝐿𝐻𝑉
𝐻𝐸
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
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 =
𝑚̇𝑒𝑡ℎ𝑎𝑛𝑜𝑙∗𝐶𝑉𝑒𝑡ℎ𝑎𝑛𝑜𝑙
𝑚̇𝑒𝑡ℎ𝑎𝑛𝑜𝑙∗𝐶𝑉𝑒𝑡ℎ𝑎𝑛𝑜𝑙 +𝑚̇𝑑𝑖𝑒𝑠𝑒𝑙∗𝐶𝑉
𝑑𝑖𝑒𝑠𝑒𝑙
Where 𝑚̇ 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
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)
𝑑𝑄𝑎𝑝𝑝𝑎𝑟𝑒𝑛𝑡 =
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
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
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
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
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
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 11.
12
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
Figure 10. Effects of lambda on NOx emissions versus spark timing [26]
Figure 11. Comparison of NOx emissions for gasoline and ethanol [26]
14
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
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
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
air cooled
4 stroke
Model
Yamaha YBR 250
Cilinder Volume (cc)
249
Compression Ratio
9.8
Bore (mm)
74
Connecting rod length (mm)
103.5
Crank radius (mm)
29
Ethanol delivery system
Direct fuel injection
Gasoline delivery system
Port injection
Engine type
17
4. Results and Discussion
4.1 Performance of SI Engine with Ethanol Direct Injection (EDI) and Gasoline
Port Fuel Injection (PFI)
4.1.1 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
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 content @ 3500 rpm
Figure 17. p-V diagrams: Different ethanol content @ 4000 rpm
19
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 content @ 4500 rpm
20
Figure 19. p-V diagrams: Different ethanol content @ 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.
4.1.2 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 26.
21
Figure 20. IMEP vs. ethanol content @ 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
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 content @ 4000 rpm
23
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
Figure 24. IMEP vs. ethanol @ 4500 rpm
Figure 25. IMEP and torque at 4500 rpm
25
Figure 26. IMEP and torque at 4500 rpm.
Figure 27. IMEP vs. ethanol content @ 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
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
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
𝑉𝑑 ∙𝑁
28
Figure 30. Zoom in on Figure 29 - Medium load.
Figure 31. BMEP vs. EER – Medium load using Equation 6.
4.1.4 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)
𝑛𝑣 =
𝑛∙𝑚𝑎
𝜌𝑎 ∙𝑉𝑑 ∙𝑁
29
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
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.
4.1.5 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
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
Figure 38. BSFC vs. EER - Medium load using Equation 8.
Figure 39. BSFC vs. EER – Medium load using Equation 8 and real RPM.
4.1.6 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
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
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
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
4.2 Ethanol Effect on Burning duration, Temperature Drop, and Compression
Ratio
4.2.1 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)
𝑃2 = 𝑃𝜃 × (
10)
𝑥𝑏 =
𝑉
𝑛
)
𝑉
𝑐
𝑃2 −𝑃1
𝑃3 −𝑃2
where:
𝑃2 :
Projection of instantaneous pressure on minimum volume line
𝑃𝜃 :
Instantaneous pressure
𝑉:
Instantaneous volume
𝑉𝑐 :
Clearance Volume
𝑛:
Polytropic index
37
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
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.
N
RPM
3500
3500
3500
3500
3500
3500
3500
3500
Sample
Code
G69 E0
G57 E4
G52.5E5.5
G49 E6.7
G45 E8
G41 E9.4
G37 E10.7
G29 E13.4
EER
%
0.00
18.66
25.58
31.03
37.01
43.23
49.15
61.11
Crank Angle Variation
Combustion Time Delay
Combustion Time Period
Δ (0-5%)
Δ (0-10%)
Δ (10-90%)
Δ (5-90%)
5.20
8.50
33.90
37.20
4.20
6.90
27.90
30.60
5.80
8.80
29.00
32.00
5.40
8.30
28.20
31.10
5.80
8.60
28.60
31.40
6.20
9.20
28.70
31.70
6.10
9.40
29.00
32.30
6.25
9.50
29.10
32.35
39
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
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 (1090%) 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.
N
RPM
4000
4000
4000
4000
4000
4000
4000
4000
Sample
Code
G69 E0
G57 E4
G52.5E5.5
G49 E6.7
G45 E8
G41 E9.4
G37 E10.7
G29 E13.4
EER
%
0.00
18.66
25.58
31.03
37.01
43.23
49.15
61.11
Crank Angle Variation
Combustion Time Delay
Combustion Time Period
Δ (0-5%)
Δ (0-10%)
Δ (10-90%)
Δ (5-90%)
5.60
8.50
30.80
33.70
5.00
7.50
29.50
32.00
5.00
7.50
30.50
33.00
6.60
9.50
30.00
32.90
6.60
9.50
28.60
31.50
7.00
9.60
25.90
28.50
7.70
10.50
27.70
30.50
8.60
11.80
30.00
33.20
41
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
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.
N
RPM
4500
4500
4500
4500
4500
4500
4500
4500
Sample
Code
G69 E0
G57 E4
G52.5E5.5
G49 E6.7
G45 E8
G41 E9.4
G37 E10.7
G29 E13.4
EER
%
0.00
18.66
25.58
31.03
37.01
43.23
49.15
61.11
Crank Angle Variation
Combustion Time Delay
Combustion Time Period
Δ (0-5%)
Δ (0-10%)
Δ (10-90%)
Δ (5-90%)
6.80
10.50
39.70
43.40
8.20
12.20
41.80
45.80
7.80
11.50
39.00
42.70
6.00
9.50
36.50
40.00
6.00
10.00
38.00
42.00
6.00
9.60
36.20
39.80
6.20
9.00
41.50
44.30
6.00
9.80
36.20
40.00
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
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
G41 E9.4
G69 E0
&
G57 E4
Figure 56. MFB curves at 5000 rpm.
Table 6. Combustion time delay and period at 5000 rpm.
N
RPM
5000
5000
5000
5000
5000
5000
5000
5000
Sample
Code
G69 E0
G57 E4
G52.5E5.5
G49 E6.7
G45 E8
G41 E9.4
G37 E10.7
G29 E13.4
EER
%
0.00
18.66
25.58
31.03
37.01
43.23
49.15
61.11
Crank Angle Variation
Combustion Time Delay
Combustion Time Period
Δ (0-5%)
Δ (0-10%)
Δ (10-90%)
Δ (5-90%)
7.80
11.80
38.80
42.80
7.90
11.00
40.50
43.60
7.50
11.50
35.70
39.70
7.00
10.50
26.50
30.00
7.50
10.10
36.40
39.00
6.10
9.80
31.40
35.10
6.00
10.00
33.80
37.80
6.20
9.90
33.10
36.80
Figure 57. Crank angle variation vs. EER at 5000 rpm: Time delay.
45
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 (05%) 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
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].
4.2.2 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
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 60.
48
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
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.
4.2.3 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 30.8. 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
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.
4.2.4 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 11.
51
Figure 63. Representation of the constant-volume cycle in the p, v and T, s diagram [34].
𝑇2 = (𝑟𝑣 )𝑘−1
11)
where:
𝑟𝑣 :
Compression Ratio
𝑘:
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 4.2.1. 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
Table 7. Theoretical temperature at start of combustion: 3500 rpm
N
RPM
3500
3500
3500
3500
3500
3500
3500
3500
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
k compression
1.3086
1.3086
1.3368
1.3649
1.3086
1.3649
1.2934
1.3659
T2
K
595.8
595.8
635.4
677.5
595.8
677.5
575.5
679.1
Table 8. Theoretical temperature at start of combustion: 4000 rpm
1
2
3
4
5
6
7
8
N
RPM
4000
4000
4000
4000
4000
4000
4000
4000
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
k compression
1.2130
1.2130
1.3368
1.3226
1.3086
1.3368
1.3649
1.3368
T2
K
479.0
479.0
635.4
615.2
595.8
635.4
677.5
635.4
Table 9. Theoretical temperature at start of combustion: 4500 rpm
1
2
3
4
5
6
7
8
N
RPM
4500
4500
4500
4500
4500
4500
4500
4500
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
k compression
1.3076
1.2814
1.2522
1.2421
1.2814
1.3236
1.3076
1.3368
T2
K
594.5
560.0
523.9
511.9
560.0
616.6
594.5
635.4
Table 10. Theoretical temperature at start of combustion: 5000 rpm
1
2
3
4
5
6
7
8
N
RPM
5000
5000
5000
5000
5000
5000
5000
5000
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
53
k compression
1.3368
1.3066
1.3368
1.3368
1.3221
1.3368
1.3368
1.3368
T2
K
635.4
593.1
635.4
635.4
614.5
635.4
635.4
635.4
The theoretical maximum temperature in the combustion process might be
obtained by Equation 12.
𝑇3 =
12)
𝐻𝑉
𝐶𝑉 ∙(𝐴⁄𝐹 +1)
+ 𝑇2
where:
𝐻𝑉:
Heating value [MJ/kg]
𝑐𝑣 :
Specific Heat at Constant Volume [J/kg-K]
𝐴⁄ : Air Fuel Ratio
𝐹
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 13.
𝐻𝑉𝑓𝑢𝑒𝑙 =
13)
̇
̇
𝑚𝑒𝑡ℎ𝑎𝑛𝑜𝑙
∙𝐻𝑉𝑒𝑡ℎ𝑎𝑛𝑜𝑙 +𝑚𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒
∙𝐻𝑉𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒
̇
̇
𝑚𝑒𝑡ℎ𝑎𝑛𝑜𝑙
+𝑚𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒
where:
𝐻𝑉𝑒𝑡ℎ𝑎𝑛𝑜𝑙 :
Heating Value Ethanol = 26.9 [MJ/kg]
𝐻𝑉𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒 :
Heating Value Gasoline = 42.9 [MJ/kg]
𝑚𝑒𝑡ℎ𝑎𝑛𝑜𝑙
̇
:
Ethanol Mass Flow [kg/h]
𝑚𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒
̇
:
Gasoline Mass Flow [kg/h]
54
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.
̇
̇
𝑚𝑒𝑡ℎ𝑎𝑛𝑜𝑙
∙𝑐𝑣 𝑒𝑡ℎ𝑎𝑛𝑜𝑙 +𝑚𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒
∙𝑐𝑣 𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒
14)
𝑐𝑣 𝑓𝑢𝑒𝑙 =
15)
𝑐𝑣 𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒 =
16)
𝑐𝑣 𝑒𝑡ℎ𝑎𝑛𝑜𝑙 =
̇
̇
𝑚𝑒𝑡ℎ𝑎𝑛𝑜𝑙
+𝑚𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒
𝑐𝑝 𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒
𝑘
𝑐𝑝 𝑒𝑡ℎ𝑎𝑛𝑜𝑙
𝑘
where:
𝑐𝑝 𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒 :
Gasoline Specific Heat at Constant Pressure = 2041 [J/kg-K]
𝑐𝑝 𝑒𝑡ℎ𝑎𝑛𝑜𝑙 :
Ethanol Specific Heat at Constant Pressure = 2339 [J/kg-K]
𝑘:
Specific Heat Ratio – Compression slope in log P vs log V
diagram
𝑚𝑔𝑎𝑠𝑜𝑙𝑖𝑛𝑒
̇
:
Gasoline Mass Flow [kg/h]
𝑚𝑒𝑡ℎ𝑎𝑛𝑜𝑙
̇
:
Ethanol Mass Flow [kg/h]
55
Table 11. Temperature at End of Combustion – 3500 rpm
1
2
3
4
5
6
7
8
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
Fuel flow G
Kg/h
0.57
0.74
0.84
0.94
1.00
1.11
1.23
1.49
Fuel Flow E
Kg/h
1.44
1.14
1.01
0.88
0.72
0.61
0.45
0.00
Total
Kg/h
2.02
1.88
1.85
1.81
1.72
1.71
1.68
1.49
HV Fuel
MJ/kg
31.46
33.20
34.13
35.16
36.22
37.24
38.61
42.90
Cv fuel
kJ/kg-K
1.722
1.698
1.649
1.601
1.655
1.573
1.640
1.494
AFR
13.10
13.42
13.49
13.61
13.43
13.88
14.10
14.46
T3
K
1891.3
1951.9
2063.5
2180.8
2112.7
2268.5
2135.1
2535.7
Table 12. Temperature at End of Combustion – 4000 rpm
1
2
3
4
5
6
7
8
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
Fuel flow G
Kg/h
0.63
0.85
0.97
1.06
1.14
1.22
1.33
1.66
Fuel Flow E
Kg/h
1.58
1.31
1.17
0.99
0.82
0.67
0.49
0.00
Total
Kg/h
2.21
2.16
2.14
2.05
1.96
1.89
1.81
1.66
HV Fuel
MJ/kg
31.46
33.20
34.13
35.16
36.22
37.24
38.61
42.90
Cv fuel
kJ/kg-K
1.858
1.832
1.649
1.652
1.655
1.606
1.554
1.527
AFR
12.80
13.20
13.23
13.41
13.85
14.09
14.26
14.45
T3
K
1706.0
1755.4
2090.0
2092.1
2069.5
2172.2
2305.6
2453.6
Table 13. Temperature at End of Combustion – 4500 rpm
1
2
3
4
5
6
7
8
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
Fuel flow G
Kg/h
0.72
0.94
1.05
1.16
1.26
1.37
1.51
1.88
Fuel Flow E
Kg/h
1.81
1.44
1.27
1.09
0.90
0.75
0.55
0.00
Total
Kg/h
2.54
2.38
2.32
2.25
2.16
2.12
2.07
1.88
HV Fuel
MJ/kg
31.46
33.20
34.13
35.16
36.22
37.24
38.61
42.90
Cv fuel
kJ/kg-K
1.724
1.734
1.760
1.759
1.690
1.622
1.622
1.527
AFR
13.18
13.18
13.26
13.38
13.80
13.99
14.22
14.14
T3
K
1881.5
1910.0
1882.9
1901.6
2007.9
2148.2
2158.9
2491.8
Table 14. Temperature at End of Combustion – 5000 rpm
1
2
3
4
5
6
7
8
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
Fuel flow G
Kg/h
0.81
1.06
1.19
1.31
1.42
1.55
1.68
2.09
Fuel Flow E
Kg/h
2.04
1.63
1.45
1.23
1.02
0.85
0.62
0.00
Total
Kg/h
2.85
2.68
2.64
2.53
2.44
2.39
2.30
2.09
56
HV Fuel
MJ/kg
31.46
33.20
34.13
35.16
36.22
37.24
38.61
42.90
Cv fuel
kJ/kg-K
1.686
1.700
1.649
1.635
1.638
1.606
1.586
1.527
AFR
13.02
13.33
13.20
13.38
13.51
14.01
14.30
14.65
T3
K
1966.57
1955.47
2092.78
2131.07
2138.16
2180.37
2226.24
2431.39
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
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
4.2.5 Estimation of Compression Ratio
The compression ratio, specific heat ratio, and thermal efficiency are related to
one another in Equation 17. In this section, the concept of thermal efficiency is
applied to mixtures of ethanol and gasoline fuels in order to estimate a compression
ratio that reaches the same thermal efficiency reached by gasoline fuel at a
determined engine speed.
1 𝑘−1
𝑛𝑡ℎ = 1 − ( )
17)
𝑟𝑣
where:
𝑛𝑡ℎ :
Thermal Efficiency [%]
𝑟𝑣 :
Compression Ratio
𝑘:
Specific Heat Ratio
The process to compute this calculation consists of two steps. First, we obtain
thermal efficiency for all the fuels tested using the real compression ratio (9.8),
and the specific heat ratio (k) calculated from the diagram log pressure versus log
volume at every EER. Second, the thermal efficiency for gasoline fuel is the
highest among the EER tested at a particular engine speed. This value becomes the
reference point to test different compression ratios in ethanol-gasoline fuel
combinations. Besides, the compression ratio and thermal efficiency have a
proportional relation. The higher the compression ratio, the greater the thermal
efficiency. Therefore, a summary of calculations is displayed in Table 15, Table
16, Table 17 and Table 18 at the engine speeds tested in this research.
59
Table 15. Estimated Compression Ratio: 3500 rpm
N
RPM
3500
3500
3500
3500
3500
3500
3500
3500
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
k compression
calculated
1.3086
1.3086
1.3368
1.3649
1.3086
1.3649
1.2934
1.3659
rv
real
9.80
9.80
9.80
9.80
9.80
9.80
9.80
9.80
nth (% )
real
50.56%
50.56%
53.64%
56.52%
50.56%
56.52%
48.81%
56.62%
rv
estimated
14.97
14.97
11.94
9.86
14.97
9.86
17.23
9.80
nth (% )
estimated
56.62%
56.62%
56.62%
56.62%
56.62%
56.62%
56.62%
56.62%
Table 16. Estimated Compression Ratio: 4000 rpm
N
RPM
4000
4000
4000
4000
4000
4000
4000
4000
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
k compression
calculated
1.2130
1.2130
1.3368
1.3226
1.3086
1.3368
1.3649
1.3368
rv
real
9.80
9.80
9.80
9.80
9.80
9.80
9.80
9.80
nth (% )
real
38.50%
38.50%
53.64%
52.11%
50.56%
53.64%
56.52%
53.64%
rv
estimated
36.94
36.92
9.80
10.84
12.07
9.80
8.22
9.80
nth (% )
estimated
53.64%
53.64%
53.64%
53.64%
53.64%
53.64%
53.64%
53.64%
Table 17. Estimated Compression Ratio: 4500 rpm
N
RPM
4500
4500
4500
4500
4500
4500
4500
4500
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
k compression
calculated
1.3076
1.2814
1.2522
1.2421
1.2814
1.3236
1.3076
1.3368
rv
real
9.80
9.80
9.80
9.80
9.80
9.80
9.80
9.80
nth (% )
real
50.44%
47.39%
43.76%
42.45%
47.39%
52.22%
50.44%
53.64%
rv
estimated
12.17
15.36
21.07
23.93
15.36
10.76
12.17
9.80
nth (% )
estimated
53.64%
53.64%
53.64%
53.64%
53.64%
53.64%
53.64%
53.64%
Table 18. Estimated Compression Ratio: 5000 rpm
N
RPM
5000
5000
5000
5000
5000
5000
5000
5000
Sample
Code
G29 E13.4
G37 E10.7
G41 E9.4
G45 E8
G49 E6.7
G52.5 E5.5
G57 E4
G69 E0
EER
%
61.1
49.1
43.2
37.0
31.0
25.6
18.7
0.0
k compression
calculated
1.3368
1.3066
1.3368
1.3368
1.3221
1.3368
1.3368
1.3368
rv
real
9.80
9.80
9.80
9.80
9.80
9.80
9.80
9.80
60
nth (% )
real
53.64%
50.33%
53.64%
53.64%
52.06%
53.64%
53.64%
53.64%
rv
estimated
9.80
12.27
9.80
9.80
10.88
9.80
9.80
9.80
nth (% )
estimated
53.64%
53.64%
53.64%
53.64%
53.64%
53.64%
53.64%
53.64%
Multiple compression ratios were presented in the tables mention previously.
For example, the higher compression ratios observed occurred at EER 18, 31, 49
and 61 when the engine speed is 3500 rpm. It means that if we want to achieve the
same thermal efficiency than gasoline fuel the compression ratio might be
increased up to 75% at this engine speed. At 4000 rpm, the higher compression
ratios are presented for EER 49 and 61. Moreover, EER 37 and 43 have reached
the larger compression ratios at 4500 rpm. Finally, the higher compression ratios
at a greater engine speed could be seen EER 49 and 31.
61
5. Conclusions
The purpose of this report was to analyze and discuss the use of ethanol as a
renewal energy in a spark ignition engine with EDI and GPI. The data obtained in
a single cylinder engine by a former Ph. D. student of the University of Technology
Sydney was used to evaluate engine performance, polytropic index, burning
duration, temperature drop and compression ratio. Speeds ranging from 3500 to
5000 rpm at medium load and EER varying from zero to sixty-one percent were
also part of these experiments. Based on the results and discussion conducted, the
following are the most significant conclusions of this report:
1. The peak pressure in the pressure-volume diagram is higher for the majority of
fuels containing ethanol than only gasoline. These maximum pressures varied
depending on the engine speed. Lower engine speeds like 3500 and 4000 rpm
tend to show more stable p-V diagrams. Fuels containing ethanol were close
to one another, and a clearance between them and the line for gasoline was
evident. Nonetheless, higher speeds such as 4500 and 5000 rpm exhibited
unbalanced p-V diagrams. In this case, the diagram corresponding to gasoline
is much closer than those containing ethanol.
2. Indicative mean effective pressure is another parameter benefited by the
injection of ethanol in the fuel. A bigger work in the expansion stroke for fuels
containing ethanol increase indicative power. Hence, the maximum pressures
recorded are those containing more ethanol at lower engine speeds; for
example, 3500 and 4000 rpm. Higher engine speeds like 4500 and 5000 rpm
do not reach larger pressures than those at reducing engine speed.
3. Brake mean effective pressure and volumetric efficiency have demonstrated
improvements when ethanol is part of the fuel depleted by the engine on
62
analysis. In contrast, the fuel consumption, carbon monoxide, and
hydrocarbons have exhibited a tendency to increase brake specific fuel
consumption and gas emissions. A rational interpretation of this performance
could be explained using the in-cylinder temperature calculated in this study.
The rise of EER reduces the maximum temperature at the end of the
combustion. This reduction leads to reducing NOx emissions, but this lower
temperature caused an overcooled region close to the cylinder wall that
augment brake specific CO and HC gas emissions.
4. The MFB curves in this paper have exhibited variation depending on engine
speed and the ethanol energy content. The combustion time delay apparently
inclines to increase with higher EERs at 3500 and 4000 rpm whereas the
opposite occurred at 4500 and 5000 rpm. This feature would be attributed to
ethanol´s lower heating value. On the other hand, the combustion time periods
have a preference to decline with EER increments at every engine speed tested.
This observation might certainly be the consequence of ethanol higher
combustion velocity or flame propagation.
5. Thought the temperature drop at the end of the combustion was addressed
before, it is important to state the effect that this measurement would have on
knock effect. The maximum pressures that can be achieved using EDI are
higher than only gasoline fuel; nevertheless, the temperature drops are
remarkable. This attribute is once again conceded to ethanol high latent heat.
6. The estimation of the compression ratio based on the temperature drop gives
an idea of the impact of high latent heat on the engine´s thermal efficiency.
However, it is believed that this calculation still needs to be analyzed in another
research so that a real understanding of this phenomenon could be explained
correctly.
63
6. Appendix
Appendix 1. Calculation of IMEP using average pressure.

Indicated Mean Effective Pressure
𝐼𝑀𝐸𝑃 =
18)
𝑃𝑖 𝑛
∀𝑑 𝑁
where:

Pi-
Indicated power [kW]
n-
Number of crank revolution in one cycle
∀𝑑 -
Volume displacement
N-
Engine speed
Indicated Power
𝑃𝑖 =
19)
𝑊.𝑁
𝑛
where:
W- Indicated work per cycle [kJ]

N-
Engine speed
n-
Number of crank revolution in one cycle
Indicated Work per Cycle
𝑊 = ∮ 𝑝 𝑑∀
20)
where:
p- Pressure
64
𝑑∀- Volume

Calculation of IMEP
The mathematical indicated mean effective pressure average (𝐼𝑀𝐸𝑃𝑎𝑣𝑒𝑟 ) for
m-conditions is expressed 21.
21)
𝐼𝑀𝐸𝑃𝑎𝑣𝑒𝑟 =
𝐼𝑀𝐸𝑃1 +𝐼𝑀𝐸𝑃2 +𝐼𝑀𝐸𝑃3 +⋯+𝐼𝑀𝐸𝑃𝑚
𝑚
The indicated mean effective pressure for condition 1 is:
22)
𝑃𝑖 1 𝑛
𝐼𝑀𝐸𝑃1 =
∀𝑑 𝑁
Replacing Equation 19 in 22 allow us to express the IMEP as a function of
indicated work per cycle and volume displacement for different conditions.
23)
𝐼𝑀𝐸𝑃1 =
24)
𝐼𝑀𝐸𝑃2 =
25)
𝐼𝑀𝐸𝑃3 =
𝑊1
∀𝑑
𝑊2
∀𝑑
𝑊3
∀𝑑
.
.
.
26)
𝐼𝑀𝐸𝑃𝑚 =
𝑊𝑚
∀𝑑
As a result, the average IMEP could be expressed as:
27)
𝐼𝑀𝐸𝑃𝑎𝑣𝑒𝑟 =
𝑊1 +𝑊2 +𝑊3 +⋯+𝑊𝑚
𝑚∗∀𝑑
65
In Equation 20, the indicated work for conditions 1, 2, 3…..,m results in:
28)
𝑊1 = ∫ 𝑝1 𝑑∀
29)
𝑊2 = ∫ 𝑝2 𝑑∀
30)
𝑊3 = ∫ 𝑝3 𝑑∀
.
.
.
𝑊𝑚 = ∫ 𝑝𝑚 𝑑∀
31)
Replacing Equation 28, 29, 30 and 31 in Equation 27give:
32)
𝐼𝑀𝐸𝑃𝑎𝑣𝑒𝑟 =
∫ 𝑝1 𝑑∀+∫ 𝑝2 𝑑∀+∫ 𝑝3 𝑑∀+⋯+∫ 𝑝𝑚 𝑑∀
𝑚∗ ∀𝑑
Because the displacement volume is constant in the cylinder, Equation 32 could be
expressed as follows:
33)
𝐼𝑀𝐸𝑃𝑎𝑣𝑒𝑟 =
∫(𝑝1 +𝑝2 +𝑝3 +⋯+𝑝𝑚 )𝑑∀
𝑚∗ ∀𝑑
If the average pressure for m conditions is defined like:
34)
𝑝𝑎𝑣𝑒𝑟 =
𝑝1 +𝑝2 +𝑝3 +⋯+𝑝𝑚
𝑚
Equation 34 could be replaced in Equation 33:
35)
𝐼𝑀𝐸𝑃𝑎𝑣𝑒𝑟 =
1
∀𝑑
∫ 𝑝𝑎𝑣𝑒𝑟 𝑑∀
66
In conclusion, the average IMEP can be calculated either by computing the IMEP
for each cycle as in 21 or by averaging the pressure as displayed in Equation 35.
67
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