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Investigation of the effect of heated ethanol fuel on combustion and

emissions of an ethanol direct injection plus gasoline port injection

(EDI+GPI) engine

Yuhan Huang *, Guang Hong

School of Electrical, Mechanical and Mechatronic Systems, University of Technology Sydney,

Sydney, Australia

Corresponding author:

Yuhan Huang, PhD

Email: Yuhan.Huang@uts.edu.au

Please cite this article as:

Y. Huang, G. Hong, Investigation of the effect of heated ethanol fuel on combustion and emissions

of an ethanol direct injection plus gasoline port injection (EDI + GPI) engine. Energy Conversion

and Management 2016; 123: 338-347. DOI: http://dx.doi.org/10.1016/j.enconman.2016.06.047

Abstract

Ethanol direct injection plus gasoline port injection (EDI+GPI) is a new technology to utilise

ethanol fuel more efficiently and flexibly in spark ignition engines. One issue needs to be addressed

in the development of EDI+GPI is the ethanol fuel's low vapour pressure and large latent heat

which slow down the ethanol's evaporation and result in the mixture unready for combustion by

the time of spark ignition and the consequent increase of CO and HC emissions. Heating the ethanol

fuel to be directly injected (EDI heating) has been proposed to address this issue. This paper reports

the investigation of the effect of EDI heating on the combustion and emissions of a research engine

equipped with EDI+GPI. The results showed that EDI heating effectively reduced the CO and HC

emissions of the engine due to the increase of evaporation rate and reduced fuel impingement and

local over-cooling. The reduction of CO and HC became more significant with the increase of

ethanol ratio. When the temperature of the ethanol fuel was increased by 40 ℃, the CO and HC

were reduced by as much as 43% and 51% respectively in EDI only condition at the original spark

timing of 15 CAD BTDC, and 15% and 47% respectively at the minimum spark advance for best

torque (MBT) timing of 19 CAD BTDC. On the other hand, the NO emission was slightly increased,

but still much smaller than that in GPI only condition due to the strong cooling effect and low

combustion temperature of EDI. The IMEP and combustion speed were slightly reduced by EDI

heating due to the decrease of injector fuel flow rate and spray collapse of flash-boiling. The largest

decrease of IMEP was 5% at the original spark timing and 3% at the MBT timing. Moreover, at

the MBT timing, the IMEP increased continuously with the increase of ethanol ratio in the entire

range from 0% to 100%. This indicated that the decrease of IMEP in high ethanol ratio conditions

at the original spark timing could be avoided by adjusting the spark timing. Therefore EDI heating

is effective to address the issues of ethanol's low evaporation rate in low temperature engine

environment and over-cooling effect at high ethanol ratio condition in the development of EDI+GPI

engine.

Keywords: Ethanol direct injection; Gasoline port injection; Heated ethanol fuel; Flash-boiling;

Combustion; Emissions

Highlights

 Effect of EDI heating on the EDI+GPI engine performance was investigated.

 CO and HC were significantly reduced and NO was slightly increased by EDI heating.

 IMEP and combustion speed were slightly reduced by EDI heating.

 EDI heating is effective to address the evaporation and over-cooling issues of EDI+GPI engine.

Abbreviations

BTDC: Before top dead centre

CAD: Crank angle degrees

EDI: Ethanol direct injection

EDI+GPI: Ethanol direct injection plus gasoline port injection

E'X': X% ethanol by volume (e.g. E40 is 40% ethanol by direct injection + 60% gasoline by port

injection)

GPI: Gasoline port injection

HRR: Heat release rate

IMEP: Indicated mean effective pressure

MBT: Minimum spark advance for best torque

SI: Spark ignition

1. Introduction

The increasing concern for the fossil fuel depletion and global warming has become the driving

force for searching renewable fuels [1]. Among the current renewable fuels, ethanol is a promising

and widely used alternative fuel for internal combustion engines [2]. Ethanol direct injection plus

gasoline port injection (EDI+GPI) is a new technology to utilise ethanol fuel more efficiently and

flexibly in spark ignition (SI) engines than E10 or E85 in the current market [3]. It integrates the

advantages of both port injection and direct injection fuel systems. Dual-injection concept offers

the flexibility to change the fuel ratio to optimize the engine performance. The GPI plays a role in

addressing the cold start issue of ethanol fuel due to ethanol's low volatility and large latent heat

of vaporization [2, 4]. It is reported that gasoline-fuelled engines could be started at ambient

temperature as low as -40 ℃ , while ethanol fuelled engines could not be started at temperature

lower than 13 ℃ without an auxiliary cold start system [5]. The EDI can be started once the engine

is warmed up and the percentage of ethanol fuel can be increased with increased engine load to

suppress the knock propensity by taking the advantages of strong cooling effect of EDI and

ethanol's high octane number . By doing so, the engine thermal efficiency can be increased by

increasing the compression ratio and using turbocharging technology whilst avoiding the knock

issue in downsized engines. The newly released Toyota D-4S engine has increased the compression

ratio to 12.7 by using a dual-injection system of gasoline port injection and direct injection in

production cars [6, 7].

Dual-injection represents an advanced combustion control strategy for both spark-ignition (e.g.

EDI+GPI for knock onset control [8, 9]) and compression-ignition (e.g. Reactivity Controlled

Compression Ignition (RCCI) for auto-ignition control [10, 11]) engines. The application of dual-

injection in SI engines was firstly proposed by Cohn et al. [12]. By direct injection of ethanol fuel

in a gasoline port injection engine, the thermal efficiency could be greatly increased by

implementing engine downsizing technologies (e.g., high compression ratio, turbocharging and

more spark advance) while avoiding the knock issue. To exploit the potential of the ethanol fuel in

increasing the compression ratio and engine thermal efficiency, a single cylinder SI engine

equipped with EDI+GPI has been developed [3]. Experimental results showed that the engine

thermal efficiency was increased and knock propensity was reduced when EDI was applied.

However, the HC and CO emissions of the engine were increased with the increase of ethanol ratio

[3, 9]. Numerical investigation to the EDI+GPI engine found that the slow evaporation rate of

ethanol fuel might have resulted in poor mixing and thus the increased HC and CO emissions [13].

Experimental and numerical studies showed that the ethanol fuel evaporated slowly when the fuel

temperature was equal to or lower than 375 K, but evaporated quickly when the temperature wa s

higher than 375 K [14, 15]. Therefore, heating the ethanol fuel to be injected directly (EDI heating)

was proposed to address the slow evaporation and poor mixing issues in the EDI+GPI engine [16].

The mixing and evaporating processes of a liquid fuel spray can be affected by a number of

factors. To achieve fast evaporation and mixing in direct injection SI engines, the most widely

adopted strategy is high pressure injection which generates very fine droplets and thus makes it

possible for the liquid fuel to evaporate in tens of crank angle degrees. However, the benefits

brought by the high pressure injection are limited. The high injection pressure may cause fuel

impingement on the cylinder and piston walls, especially in downsized engines. On the other hand,

the droplet size reduced by increasing injection pressure is limited by the injection pressure in a

medium range [17]. Fuel temperature is also an important factor that affects the droplet size and

evaporation rate of a liquid fuel. Increasing the fuel temperature is an effective way to reduce the

viscosity and increase the vapour pressure of the fuel, and consequently increase the break-up and

evaporation rates. Since ethanol fuel has larger viscosity and surface tension than that of gasoline,

E85 shows larger Sauter Mean Diameter (SMD) than gasoline does. With the increase of fuel

temperature, the SMD reduces and the difference in E85 and gasoline's SMDs becomes small when

the temperature is sufficiently high [18]. Particularly, when the fuel temperature is higher than its

boiling point, vapour bubbles may form inside a droplet and burst into smaller ones [19]. It is

reported that the SMD of the spray droplets showed a rapid reduction when flash-boiling occurred

[20]. Moreover, the vapour pressure of ethanol fuel is lower than that of gasoline fuel when the

fuel temperature is lower than 375 K but higher when the fuel temperature is higher than 375 K.

As a result, ethanol evaporates more slowly than gasoline does in low temperature condition, but

they reach the similar evaporation rates when the fuel temperature is higher than 375 K [14]. Xu.

et al. [21] proposed using flash-boiling to generate fine droplets as an alternative and economic

method for high pressure injection. Investigation showed that the evaporation rate of a spray could

be significantly enhanced in flash-boiling conditions [14, 17].

Nwafor [22, 23] investigated the combustion and emission performance of a diesel engine

fuelled with elevated fuel temperature of neat vegetable oil. The effect of fuel temperature on the

performance of a diesel engine fuelled with rapeseed methyl ester (RME) fuel was investigated

[24]. Kabasin et al. [25, 26] investigated the cold start performance of ethanol-fuelled SI engines

equipped with heated injectors. The engine cold start time was reduced and the HC and CO

emissions were significantly decreased when the fuel was heated. Sales and Sodré [5, 27]

investigated the cold start combustion and emission characteristics of a flexible fuel engine with

heated intake air and ethanol fuel. By heating the intake air and the ethanol fuel, their results

showed that the ethanol fuelled engine could be started at cold ambient temperature of 0 ℃ and

with significant reduction in HC and CO emissions.

As reviewed above, increasing the fuel temperature could be an effective and economic way

to generate fine and fast evaporation sprays and to improve the performance of ethanol fuelled

engines in cold start conditions. However, so far, rare work has been reported on investigating the

engine performance with heated ethanol fuel at warmed up conditions. To address the emission

issues in the current EDI+GPI engine, experiments were conducted to investigate the effect of EDI

heating on improving the combustion and emissions of an SI engine equipped with EDI+GPI dual-

injection system.

2. Experimental Apparatus and Procedures

2.1. EDI+GPI engine test rig

The experiments were conducted on a single-cylinder four-stroke air-cooled SI engine

equipped with a direct injection system for injecting ethanol fuel and a port injection system for

injecting gasoline fuel. Fig. 1 shows the schematic of the engine test rig. Table 1 provides the

specifications of the engine which was originally a port fuel injection engine used on the Yamaha

YBR250 motorcycle. The engine was modified to be an EDI+GPI engine by adding an EDI fuel

system to the engine. The EDI injector was a six-hole injector with a nozzle diameter of 110 um.

The spray cone angle is 34° and the bent angle is 17°. The GPI injector was the original port fuel

injector used in the Yamaha YBR250 motorcycle. The EDI injector was mounted with the spray

plumes bent towards the spark plug to create an ignitable mixture around the spark plug. The tip of

the injector was placed 15 mm to the spark plug on the intake valve side. Both the GPI injector and

EDI injector were controlled by an electronic control unit (ECU). The EDI+GPI fuel system offers

the flexibility to operate the engine over a full range of ethanol ratio from 0% (GPI only) to 100%

(EDI only). More information about the engine test rig can be found in [3, 19].

Table 1 Specifications of the EDI+GPI engine.

Single cylinder, air cooled, four-stroke

An electric resistance heater made of Kanthal A1 heating resistance wire was used to heat the

ethanol fuel in the high pressure fuel rail, as shown in Fig. 1. The wire was wrapped on the fuel rail

upstream the injector. A T-type thermocouple was attached on the surface of the fuel rail at the

entrance of the EDI injector to measure the ethanol fuel temperature. The temperature was fed back

to a 2132 Eurotherm PID temperature controller which controlled the relay of the heating system

so that the heating process was stopped when the fuel temperature reached the required value. The

fuel rail temperature at the injector entrance was regarded as the fuel temperature in the present

study. The accuracy of the temperature control was within ±3 ℃.

Fig. 1. Schematic of the EDI+GPI engine.

The mass flow rate of the intake air was measured with a ToCeil-20N hot-wire thermal flow

meter. The engine speed was controlled by an eddy current DC dynamometer. The in-cylinder

pressure was measured by a Kistler 6115B measuring spark plug pressure transducer at a resolution

of 0.5 crank angle degree. A hundred consecutive cycles of pressure were recorded. The exhaust

gas emissions were measured using a Horiba MEXA-584 L gas analyzer. The Horiba MEXA-584L

gas analyzer can measure the lambda (excess air ratio) of multiple fuels with H/C and O/C atomic

ratios of the fuel input by the user. In the present study, the lambda was monitored and kept around

one by adjusting the mass flow rates of the gasoline and ethanol fuels at a designated fuel ratio and

a fixed throttle position. The gasoline and ethanol fuel mass flow rates were input to the ECU. The

calibration of the EDI injector was provided by the ECU supplier, Hents Technologies Inc. The

EDI injection pressure was controlled by the ECU and the solenoid valve in the high pressure pump.

2.2. Test fuels

Table 2 provides the main properties of gasoline and ethanol fuels used in the present study.

The gasoline fuel was the Unleaded Petrol (ULP) from the Caltex Australia with an octane number

of 91. The ethanol fuel wa s provided by the Manildra Group.

Table 2 Properties of ethanol and gasoline fuels at 300 K.

Research octane number (-)

Stoichiometric air/fuel ratio (-)

Lower heating value (MJ/kg)

Enthalpy of vaporization (kJ/kg)

Saturation vapor pressure (kPa)

Laminar flame velocity @ λ=1,

100 kPa, 100 ℃ (m/s)

2.3. Experimental procedures

The engine was started and warmed up by gasoline port injection only. The experiments were

conducted when the engine body temperature became stable at around 200 ℃. Then the amount of

gasoline was reduced and the ethanol fuel wa s directly injected into the cylinder to substitute

gasoline. Table 3 lists the engine conditions investigated in the present study. The experiments

were conducted at medium load (IMEP 6.0~6.5 bar). The tested engine speeds were 3500 and 4000

rpm. In all the tests, the lambda was kept around one. The GPI pressure was 0.25 MPa and the EDI

pressure wa s 4 MPa. The GPI timing was 410 CAD BTDC and EDI timing was 300 CAD BTDC.

The ethanol ratio started at E0 and then varied from E25 to E100 with an increment of 15% (E ' X'

means X% ethanol by volume. e.g. E40 means 40% ethanol by direct injection plus 60% gasoline

by port injection). Firstly the experiments were conducted to compare the engine performance with

and without EDI heating at the engine original spark timing of 15 CAD BTDC. The results showed

significant reduction of engine emissions but slightly decreased indicated mean effective pressure

(IMEP). To recover the IMEP, the effect of EDI heating on engine combustion and emissions was

also investigated at the minimum spark advance for best torque (MBT) timing of 19 CAD BTDC.

The experiments were firstly performed without the EDI heating system. The measured ethanol

temperature was around 50℃ (± 2℃ ). Then the EDI heating system was applied to heat the ethanol

fuel to 70 ℃ and 90 ℃. The power required for EDI heating was dependent on the fuel flow rate

and the heating temperature. The maximum heating power was about 150 W at E100, 90 ℃ and

4000 rpm condition, which was about 3% of the engine output power. However, the heater only

worked intermittently at this power and the power required was much smaller in lower ethanol ratio

conditions. In production engines, the heating energy can be taken from the high-temperature

exhaust gas (~400 ℃ in the present study) and thus eliminate the parasitic load of the heating

element in the present experiments.

Table 3 Experimental conditions.

0%, 25%, 40%, 55%, 70%, 85%, 100%

15 (original), 19 (MBT) CAD BTDC

50 (baseline), 70 and 90 ℃

To ensure the accuracy of the results, five samples were recorded at each tested engine

condition. The maximum standard deviation of the measurements were 5% for ISCO, 3% for ISNO,

5% for ISHC and 1% for lambda, which indicated acceptable quality of the experimental data and

justified the use of averaged sample data in the present study.

3. Results and Discussion

The experimental results will be presented and discussed as follows. Sections 3.1 reports

the effect of EDI heating on the engine performance represented by IMEP and emissions of CO,

HC and NO at the original engine spark timing of 15 CAD BTDC. In Section 3.2, the combustion

characteristics will be discussed to analyse the causes of the results at the original spark timing.

Section 3.3 will focus on the enhanced effect of EDI heating by adjusting the spark timing to the

MBT timing of 19 CAD BTDC.

3.1. Effect of EDI heating on engine performance at the original spark timing

Fig. 2 shows the effect of ethanol fuel temperature on the IMEP of the EDI+GPI engine at

different ethanol ratios and engine speeds. The spark timing was 15 CAD BTDC which was the

timing set for GPI only in the original engine control unit. As shown in Fig. 2, the IMEP increases

with the increase of ethanol ratio from 0% to 70% without EDI heating, resulting from the effective

charge cooling effect and faster combustion speed of ethanol fuel. However, the IMEP decreases

when the ethanol ratio is further increased from 70% to 100%. This is because the mixture becomes

over-lean around the spark plug, and local over-cooling and severe fuel impingement occur at high

ethanol ratios [28]. When the ethanol fuel temperature is increased from 50 ℃ to 70 ℃, the IMEP

decreases slightly at each ethanol ratio. It decreases noticeably at high ethanol ratios when the

ethanol fuel temperature reaches 90 ℃. The maximum reduction of IMEP at 3500 rpm is 2.2% at

fuel temperature of 70 ℃ and 4.5% at 90 ℃ when the ethanol ratio is 70%. The maximum reduction

of IMEP at 4000 rpm is 2.2% at 70 ℃ and 5.4% at 90 ℃ at the ethanol ratio of 70%. This is caused

by the decrease of the combustion speed when the ethanol fuel is heated, which will be discussed

referring to Fig. 7. Fig. 3 shows the effect of ethanol fuel temperature on the indicated thermal

efficiency. In the present study, the lambda was kept around one. As the air-fuel ratio of ethanol is

9.0 and that of gasoline is 14.8, 64% more mass of ethanol is injected to replace the reduced mass

of gasoline fuel in order to keep the lambda at one. On the other hand, the heating value of gasoline

fuel (42.9 MJ/kg) is 59% higher than that of ethanol fuel (26.9 MJ/kg). This indicates that the total

energy input at each ethanol ratio condition is close to that of GPI only condition. Consequently,

the effect of ethanol fuel temperature on the indicated thermal efficiency shows the same tendencies

as that of IMEP in Fig. 2. As shown in Fig. 3, the thermal efficiency increases with the increase of

ethanol ratio from 0% to 70% without EDI heating, but decreases when the ethanol ratio is further

increased from 70% to 100%. When EDI heating is applied, the thermal efficiency decreases

slightly at ethanol fuel temperature of 70 ℃ and decreases noticeably at 90 ℃.

Fig. 2. Effect of ethanol fuel temperature on the IMEP.

Fig. 3. Effect of ethanol fuel temperature on the indicated thermal efficiency.

5.8

6.0

6.2

6.4

6.6

0% 20% 40% 60% 80% 100%

IMEP (bar)

Ethanol ratio by volume (%)

3500rpm 50℃(No EDI heating)

3500rpm 70℃

3500rpm 90℃

4000rpm 50℃(No EDI heating)

4000rpm 70℃

4000rpm 90℃

25%

26%

27%

28%

0% 20% 40% 60% 80% 100%

Indicated thermal efficiency (%)

Ethanol ratio by volume (%)

3500rpm 50℃(No EDI heating)

3500rpm 70℃

3500rpm 90℃

4000rpm 50℃(No EDI heating)

4000rpm 70℃

4000rpm 90℃

Figs. 4 and 5 show significant reduction of indicated specific carbon monoxide (ISCO) and

hydrocarbon (ISHC) emissions with EDI heating. As shown in Figs. 4 and 5, without EDI heating,

both the ISCO and ISHC emissions increase with the increase of ethanol ratio due to the incomplete

combustion caused by ethanol's poor evaporation, local over-cooling and fuel impingement of EDI

[28]. When the ethanol fuel is heated, both ISCO and ISHC emissions are reduced significantly,

and this reduction is enhanced with the increase of ethanol ratio. Compared with that without EDI

heating, as shown in Fig. 4, the ISCO emission is reduced by 21.8%-43.2% at 3500 rpm and 11.1%-

30.6% at 4000 rpm when the ethanol fuel temperature is increased from 50 ℃ to 70 ℃ in the

ethanol ratio range of 25%-100%. The ISCO is reduced by 39.2%-49.1% at 3500 rpm and 41.0%-

47.7% at 4000 rpm when the fuel temperature is increased from 50 ℃ to 90 ℃. As shown in Fig.

5, compared with that without EDI heating, the ISHC emission is reduced by 10.8%-29.9% at 3500

rpm and 14.1%-46.0% at 4000 rpm when the ethanol fuel temperature is increased from 50 ℃ to

70 ℃. The ISHC is reduced by 15.7%-38.1% at 3500 rpm and 20.6%-61.2% at 4000 rpm when

the fuel temperature is increased from 50 ℃ to 90 ℃. Particularly, in low ethanol ratio conditions,

EDI heating effectively reduces the CO (25%-40% ethanol ratio) and HC (25%-70% ethanol ratio)

emissions to be lower than that in GPI only condition. The reduction of ISCO and ISHC emissions

can be attributed to the improved ethanol fuel's evaporation and mixing processes which were the

original aims of EDI heating. As analysed in the numerical investigation of the same engine used

in the present study [13, 28], the low evaporation rate of ethanol fuel resulted in a large number of

ethanol droplets in the late compression and combustion strokes. The in-cylinder flows were slow

and thus the heat and mass transfer between the liquid fuel and air were weak. This caused not only

lean mixture around the spark plug, but over-rich mixture and over-cooling in the regions where

the ethanol droplets were concentrated in. Moreover, the fuel impingement on the cylinder and

piston walls became severe when the ethanol ratio was high, due to the longer spray penetration

length of longer injection duration. The uneven distribution of mixture, local over-cooling effect,

unburnt liquid ethanol droplets and fuel impingement caused the increase of ISCO and ISHC

emissions with the increase of ethanol ratio in the engine conditions without EDI heating. However,

when EDI heating is applied, the evaporation of ethanol fuel should be accelerated. As a result, the

liquid ethanol droplets, local over-cooling and fuel impingement during combustion process should

be reduced, resulting in the reduction of ISCO and ISHC emissions. Fig. 9 shows the images taken

from the ethanol fuel injected in a constant volume chamber. It shows that the EDI spray tip

penetration is slightly decreased due to the enhanced evaporation and spray-air interaction when

the ethanol fuel is heated from 52 ℃ to 92 ℃. Moreover, more fuel in the spray will be in vapour

phase when the fuel is heated than that without fuel heating. Fuel impingement will not occur when

vapour fuel reaches the cylinder wall. Significant reduction of CO and HC emissions was also

reported in the investigation of pre-heating the ethanol fuel in cold start conditions [25-27, 29].

Fig. 4. Variation of ISCO with ethanol fuel temperature.

Fig. 5. Variation of ISHC with ethanol fuel temperature.

Fig. 6 shows the effect of EDI heating on indicated specific nitric oxide (ISNO) emission. As

shown in Fig. 6, the influence of EDI heating on ISNO is insignificant. With or without EDI heating,

the ISNO emission reduces quickly with the increase of ethanol ratio due to the strong cooling

effect of EDI and lower combustion temperature of ethanol fuel [13, 30]. When EDI heating is

applied, the ISNO emission increases slightly at each ethanol ratio condition. This may be caused

by the increased combustion temperature resulted from the heated fuel. However, even with EDI

heating, the ISNO emission is still much smaller than that of GPI only condition. The slight increase

of NO emission by EDI heating was also observed for ethanol fuelled engines in cold start

conditions [25, 26, 31].

0% 20% 40% 60% 80% 100%

ISCO (g/kw-h)

Ethanol ratio by volume (%)

3500rpm 50℃(No EDI heating)

3500rpm 70℃

3500rpm 90℃

4000rpm 50℃(No EDI heating)

4000rpm 70℃

4000rpm 90℃

0% 20% 40% 60% 80% 100%

ISHC (g/kw-h)

Ethanol ratio by volume (%)

3500rpm 50℃(No EDI heating)

3500rpm 70℃

3500rpm 90℃

4000rpm 50℃(No EDI heating)

4000rpm 70℃

4000rpm 90℃

Fig. 6. Variation of ISNO with ethanol fuel temperature.

3.2. Combustion characteristics with EDI heating at the original spark timing

Fig. 7 shows the in-cylinder pressure and the corresponding heat release rate (HRR) at the

ethanol fuel temperatures of 50 ℃ (no fuel heating), 70 ℃ and 90 ℃. The engine speed is 3500

rpm and the ethanol ratio is 70%. As shown in Fig. 7, the starting phase of HRR is not affected by

the ethanol fuel temperature. However, HRR reduces with the increased fuel temperature after that,

indicating slower combustion speed. As a result, the peak in-cylinder pressure is reduced during

the combustion process with the increase of ethanol fuel temperature. This reduced peak pressure

was also observed in compression-ignition engines fuelled with heated rapeseed methyl ester (RME)

[24], dimethyl ether (DME) [32], diesel and cetane [33].

Fig. 7. Effect of ethanol fuel temperature on the in-cylinder pressure and heat release rate.

0% 20% 40% 60% 80% 100%

ISNO (g/kw-h)

Ethanol ratio by volume (%)

3500rpm 50℃(No EDI heating)

3500rpm 70℃

3500rpm 90℃

4000rpm 50℃(No EDI heating)

4000rpm 70℃

4000rpm 90℃

-5

340 360 380 400 420

Heat release rate (J/CAD)

In-cylinder pressure (bar)

Crank angle degrees (ATDC)

E70 50°C (No EDI heating)

E70 70°C

E70 90°C

The decrease of combustion speed with increased ethanol fuel temperature may be caused by

the following two causes. Firstly, the fuel viscosity, surface tension and density decreasing with

the increased fuel temperature should have affected the injection process. When the ethanol fuel

temperature is increased from 50 ℃ to 90 ℃, the density of the ethanol fuel decreases from 763 to

721 kg/m3, viscosity decreases from 0.676 to 0.374 mPa-s, and surface tension decreases from

0.0202 to 0.0152 N/m [34]. Experiments and simulations on the fuel injection process showed that,

when the fuel temperature was increased, the actual injection timing was retarded, the peak rail

pressure was decreased and the injection duration was prolonged [32, 35]. The fuel mass flow rate

was decreased with the increase of fuel temperature because of the decreased fuel density [24, 36,

37]. Furthermore, the size of vapour bubbles at the injector's nozzle exit increased with the increase

of the fuel temperature [38], which would decrease the nozzle discharge coefficient. All these

factors reduced the mass of ethanol fuel directly injected into the combustion chamber when the

ethanol was heated in the present study. This has been evidenced by the slight increase of lambda

with the increase of fuel temperature in the experiments. Fig. 8 shows the variation of measured

lambda values with ethanol fuel temperature at different ethanol ratios. As described in Section 2.3,

to compare the results with and without EDI heating, the throttle opening, injection pressure and

pulse width were kept unchanged at each engine speed and ethanol ratio. So that the intake airflow

rates were the same in both test conditions with and without EDI heating. As shown in Fig. 8, the

lambda increases slightly with the increase of fuel temperature at each ethanol ratio and the change

becomes more significant at high ethanol ratios. At ethanol ratio of 100%, the lambda increases by

about 2% for every 20 ℃ of fuel temperature increase at the two tested engine speeds. The

reduction of ethanol fuel supply might lead to the decrease of IMEP by EDI heating. This suggests

that the injection pulse width should be increased to get the lambda back to one when implementing

EDI heating in order to avoid the decrease of IMEP.

Fig. 8. Variation of lambda with ethanol fuel temperature.

The second cause to the reduced IMEP is that the fuel spray process might be changed by

flash-boiling which is a phenomenon occurs when the fuel temperature is higher than the saturation

0.94

0.97

1.00

1.03

1.06

0% 20% 40% 60% 80% 100%

Measured lambda (-)

Ethanol ratio by volume (%)

3500rpm 50℃(No EDI heating)

3500rpm 70℃

3500rpm 90℃

0.94

0.97

1.00

1.03

1.06

0% 20% 40% 60% 80% 100%

Measured lambda (-)

Ethanol ratio by volume (%)

4000rpm 50℃(No EDI heating)

4000rpm 70℃

4000rpm 90℃

temperature of the fuel at the ambient pressure. Experiments showed that the multi-jet spray might

collapse to a single-jet spray when flash-boiling occurred by either increasing the fuel temperature

or decreasing the ambient pressure [18, 20, 39]. The EDI injector used in the present study was a

six-hole nozzle. Fig. 9 shows the spray structures of the EDI injector at 2.0 ms after the start of

injection in a constant volume chamber at three ethanol fuel temperatures of 52℃, 77 ℃ and 92 ℃

[19]. The injection pressure was 60 bar and the ambient pressure was 1 bar, which reproduced the

in-cylinder conditions of early EDI injection in the present engine experiments. As shown in Fig.

9, the two side plumes of the ethanol spray converge towards the middle one when the fuel

temperature is increased from 52 ℃ to 77 ℃ . The three spray plumes collapse completely when the

temperature reaches 92 ℃. As described in Section 2.1, the EDI injector has a bent angle of 17°

and was installed with the spray plumes bent towards the spark plug in the EDI+GPI engine to

create an ignitable mixture around the plug gap by the time of ignition. However, the results in Fig.

9 showed that the spray was not collapsed at 52 ℃ or 77 ℃, but collapsed only at 92 ℃. The

collapsed spray might deform the designed fuel distribution in the combustion chamber, and

consequently slow down the ignition and combustion processes, resulting in the reduced peak

cylinder pressure and heat release rate as shown in Fig. 7.

(a) 52 ℃ (b) 77 ℃ (c) 92 ℃

Fig. 9. Effect fuel temperature on the EDI spray structure [19].

Fig. 10 shows the variation of combustion initiation duration with ethanol fuel temperature at

different ethanol ratios at the original spark timing of 15 CAD BTDC. The combustion initiation

duration, denoted by CA0-10%, is defined as the crank angle degrees from the spark timing to the

timing of 10% of the fuel mass fraction burnt. CA0-10% is directly related to the combustion

stability. Shorter CA0-10% means more stable combustion [40]. As shown in Fig. 10, without EDI

heating, the CA0-10% at ethanol ratio in the range of 25%-70% is shorter than that in GPI only

condition. This indicates an improved combustion stability with ethanol ratio up to 70%. It can be

attributed to the faster flame speed of ethanol fuel. However, the CA0-10% increases when the

ethanol ratio is higher than 70%. This may be because the mixture wa s too lean due to the low

evaporation rate of ethanol fuel and local over-cooling and fuel impingement occurred in high

ethanol ratio conditions [28]. As shown in Fig. 10, the ethanol ratio of 70% at which the CA0-10%

starts to increase has been extended to 85% with EDI heating. With EDI heating, the CA0-10% is

shorter than that without EDI heating when the ethanol ratio is greater than 70%. This may be

because EDI heating supplies additional thermal energy required by the ethanol droplets to

evaporate, which reduces the over-cooling problem in higher ethanol ratio conditions. However,

the CA0-10% with EDI heating is longer than that without EDI heating when the ethanol ratio is

less than 70%. This may be due to the reduced mass of ethanol fuel injected and spray collapse, as

discussed for the results in Figs. 8 and 9.

Fig. 10. Variation of CA0-10% with ethanol fuel temperature.

Fig. 11 shows the effect of ethanol fuel temperature on the major combustion duration at

different ethanol ratios at the original spark timing of 15 CAD BTDC. The major combustion

duration, denoted by CA10-90%, is defined as the crank angle degrees from 10% to 90% of the

fuel mass fraction burnt. The shorter the CA10-90% is, the closer the combustion process is to the

constant volume process which consequently results in higher thermal efficiency [40]. As shown

in Fig. 11, in general, the CA10-90% is longer with EDI heating than that without EDI heatnig.

Consistently with the causes to the longer initial combustion duration, the combustion speed is

decreased by the the reduced mass of ethanol fuel and the deformation of ethanol spray structure

when EDI heating is applied. Less mass of ethanol fuel injected makes the mixture leaner which

slows down the combustion speed. Moreover the deformation of ethanol spray changes the

designed in-cylinder mixture distribution and consequently results in longer combustion duration.

Fig. 11. Variation of CA10-90% with ethanol fuel temperature.

0% 20% 40% 60% 80% 100%

CA0-10 (CAD)

Ethanol ratio by volume (%)

3500rpm 50℃(No EDI Heating)

3500rpm 70℃

3500rpm 90℃

0% 20% 40% 60% 80% 100%

CA0-10 (CAD)

Ethanol ratio by volume (%)

4000rpm 50℃(No EDI Heating)

4000rpm 70℃

4000rpm 90℃

0% 20% 40% 60% 80% 100%

CA10-90 (CAD)

Ethanol ratio by volume (%)

3500rpm 50℃(No EDI Heating)

3500rpm 70℃

3500rpm 90℃

0% 20% 40% 60% 80% 100%

CA10-90 (CAD)

Ethanol ratio by volume (%)

4000rpm 50℃(No EDI Heating)

4000rpm 70℃

4000rpm 90℃

3.3. Effect of EDI heating on engine performance and emissions at the MBT timing

As shown in Figs. 4 and 5, the CO and HC emissions were reduced effectively by EDI heating.

However, the IMEP was reduced when EDI heating was applied. As anaylsed in 3.2, the decrease

of IMEP may be due to the change in combustion phase and the associated decrease of combustion

speed. To recover the IMEP, experiments were conducted to find the 'best' spark timing when EDI

heating was applied to the EDI+GPI engine. This section reports the effect of EDI heating on the

engine performance at the minimum spark advance for best torque (MBT) timing.

Fig. 12 shows the variations of IMEP and NO emission with spark timing. The engine speed

is 3500 rpm and the three ethanol ratios are 0% (GPI only), 55% (EDI+GPI) and 100% (EDI only).

As shown in Fig. 12(a), the IMEP increases quickly with the advance of spark timing from 15 to

22 CAD BTDC. It becomes stable at the spark timing of 22 to 24 CAD BTDC and then decreases

with further advance of spark timing. The maximum IMEP can be achieved at the spark timing

around 23 CAD BTDC for the designated three ethanol ratios. As shown in Fig. 12(b), the NO

emission increases sharply with the advance of spark timing. This suggests that if the spark timing

is slightly retarded from the maximum IMEP timing, the engine output power would hardly suffer

but the NO emission would be significantly reduced. Moreover, slight knock occurs when the spark

timing is earlier than 24 CAD BTDC in GPI only condition. Therefore, the MBT timing is defined

as a spark retard of 4 degrees from the angle of maximum IMEP [41]. Based on this definition, the

MBT timing is 19 CAD BTDC in the present study.

(a) (b)

Fig. 12. Effect of spark timing on IMEP (a) and NO emission (b) at ethanol ratios of 0%, 55%

and 100%.

Fig. 13 shows the variation of IMEP with ethanol fuel temperature. The spark timing is the

MBT timing and the engine speed is 3500 rpm. As shown in Fig. 13, the IMEP increases

continuously with the increase of ethanol ratio in the entire range from 0% to 100%. This implies

5.8

6.0

6.2

6.4

6.6

14 16 18 20 22 24 26 28 30 32

IMEP (bar)

Spark timing (CAD BTDC)

E55

E100

14 16 18 20 22 24 26 28 30 32

ISNO (g/kw-h)

Spark timing (CAD BTDC)

E55

E100

that the decrease of IMEP in high ethanol ratio conditions (as shown in Fig. 2) can be avoided by

adjusting the spark timing. When the ethanol fuel temperature is increased from 50 ℃ to 70 ℃, the

IMEP is lower than that without EDI heating but still increases with the increase of ethanol ratio.

However, IMEP becomes to decrease with the ethanol ratio and noticeably lower than that at 50 ℃

and 70 ℃ in high ethanol ratio conditions when the fuel temperature is further increased to 90 ℃.

This may be caused by the decrease of fuel supply and collapse of spray when the fuel temperature

is too high like 90 ℃, as discussed in Section 3.2. The maximum reduction in IMEP is 0.8% at

ethanol fuel temperature of 70 ℃ and 3.0% at 90 ℃ at the ethanol ratio of 85% which is much

smaller than that at the original spark timing of 15 CAD BTDC.

Fig. 13. Effect of EDI heating on IMEP at the MBT timing at various ethanol ratios.

Fig. 14 shows the effect of EDI heating on the emissions of ISCO, ISHC and ISNO at the MBT

spark timing. Comparison based on spark timing shows that the CO, HC and NO emissions at MBT

timing are slightly higher than that at the original spark timing of 15 CAD BTDC. However,

compared with that without EDI heating, both ISCO and ISHC emissions at MBT timing are

significantly reduced and this reduction is enhanced by the increase of ethanol fuel temperature.

The maximum reduction is 15% for ISCO and 47% for ISHC at E100 when the ethanol fuel

temperature is increased from 50 ℃ to 90 ℃ . Particularly, the ISHC emission reduces significantly

in the ethanol ratio range of 70%-100% at the ethanol fuel temperature of 90℃. Meanwhile, the

ISNO emission is slightly increased by EDI heating, as shown in Fig. 14(c). However this increase

does not offset the ISNO reduction caused by the strong cooling effect and low combustion

temperature of EDI.

5.8

6.0

6.2

6.4

6.6

0% 20% 40% 60% 80% 100%

IMEP (bar)

Ethanol ratio by volume (%)

50℃ MBT (No EDI heating) 50℃ Original

70℃ MBT 70℃ Original

90℃ MBT 90℃ Original

Fig. 14. Effect of EDI heating on ISCO (a), ISHC (b) and ISNO (c) at the MBT timing.

0% 20% 40% 60% 80% 100%

ISCO (g/kw-h)

Ethanol ratio by volume (%)

50℃ MBT (No EDI heating)

70℃ MBT

90℃ MBT

50℃ Original

70℃ Original

90℃ Original

0% 20% 40% 60% 80% 100%

ISHC (g/kw-h

Ethanol ratio by volume (%)

50℃ MBT (No EDI heating)

70℃ MBT

90℃ MBT

50℃ Original

70℃ Original

90℃ Original

0% 20% 40% 60% 80% 100%

ISNO (g/kw-h)

Ethanol ratol by volume (%)

50℃ MBT (No EDI heating)

70℃ MBT

90℃ MBT

50℃ Original

70℃ Original

90℃ Original

4. Summary and Conclusions

Experiments were conducted to investigate the effect of heated ethanol fuel on improving the

combustion and emission performance of a single-cylinder SI engine equipped with EDI+GPI. The

tested engine was operated at medium load at the engine speeds of 3500 rpm and 4000 rpm. The

volume ratio of the ethanol fuel in direct injection was varied from 0% (GPI only) to 100% (EDI

only). The lambda was around one. The ethanol fuel temperature was increased from 50 ℃ (no

EDI heating) to 70 ℃ and 90 ℃ (flash-boiling spray). The effect of EDI heating on the engine

performance was investigated at both the original spark timing of 15 CAD BTDC and the MBT

timing of 19 CAD BTDC. The conclusions of this study can be drawn as follows.

(1) At the original engine's spark timing of 15 CAD BTDC, EDI heating effectively reduced the

ISCO and ISHC emissions due to increased ethanol's evaporation rate, and reduced fuel

impingement and local over-cooling effect. The reduction of IS CO and ISHC was enhanced

with the increase of ethanol ratio. When the ethanol fuel was heated from 50 ℃ to 90 ℃, the

ISCO was reduced by 43% and the ISHC was reduced by 51% in EDI only condition. On the

other hand, the ISNO emission was slightly increased, but still much smaller than that in GPI

only condition due to EDI's strong cooling effect and low combustion temperature. However,

the IMEP and combustion speed were slightly reduced by EDI heating. The largest decrease

of IMEP was about 5% at the ethanol ratio of 70% when ethanol fuel temperature was

increased from 50 ℃ to 90 ℃.

(2) At the MBT spark timing of 19 CAD BTDC, the IMEP increased with the increase of ethanol

ratio in the entire range from 0% to 100%. This indicated that the decrease of IMEP in high

ethanol ratio conditions at the original spark timing of 15 CAD BTDC was avoided by

adjusting the spark timing. When the ethanol fuel temperature was increased from 50 ℃ to

90 ℃, the ISCO was reduced by 15% and ISHC was reduced by 47% in EDI only condition.

Meanwhile, the reduction of IMEP by EDI heating was less than 3% and smaller than that at

the original spark timing.

(3) When the ethanol fuel temperature was increased to the flash-boiling region, the ethanol spray

collapsed which deformed the designed fuel distribution in the combustion chamber and

consequently deteriorated the combustion process. Meanwhile, the mass of fuel injected into

the cylinder decreased when the ethanol fuel was heated, resulted from the decreased fuel

density, viscosity, surface tension and injector discharge coefficient. As a result, the IMEP and

combustion speed were slightly decreased with EDI heating.

(4) Results indicate that EDI heating is effective to address the issues of ethanol's low evaporation

rate in low temperature engine environment and over-cooling effect at high ethanol ratio

condition in the development of EDI+GPI engines in terms of minimizing the engine emissions.

Acknowledgement

The scholarship provided by the China Scholarship Council (CSC) is gratefully appreciated.

The authors would like to express their great appreciation to Manildra Group for providing the

ethanol fuel.

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... Sales and Sodré [45] observed substantial reductions in HC and CO emissions in a port-fuel injection (PFI) engine when using heated ethanol and air intake compared to unheated gasoline injection, which is a traditional coldstart system used in Brazil. Huang and Hong [19] found a decrease of 43% and 51% in CO and HC emissions, respectively, when heating ethanol by 40 • C in a direct inject (DI) engine, although they observed a slight increase in NO emission. In a DI single-cylinder test using heated gasoline, Fedor et al. [12] reached 24% and 53% reductions in HC and particulate emissions, while in a DI engine they found reductions of 27% and 8% for the same pollutants. ...

... In a DI single-cylinder test using heated gasoline, Fedor et al. [12] reached 24% and 53% reductions in HC and particulate emissions, while in a DI engine they found reductions of 27% and 8% for the same pollutants. These results are possible because heating improves the fuel atomization [11,20] and, in the case of DI engines, increases fuel evaporation and reduces over-cooling problems, especially with higher ethanol ratios in the injected fuel [19]. This allows a reduction in the mass injected into the combustion chamber and, consequently, a reduction in pollutant emissions. ...

Heating fuels before injection into internal combustion engines provides benefits related to the reduction of pollutant emissions and the improvement in cold-start performance. Current fuel-heating solutions used in cold-start systems of ethanol-fueled engines in Brazil involve nucleate boiling heat transfer and, in order to extend its use to fossil fuels, further studies are required on the heating process of multicomponent mixtures. Because fuels like gasoline, diesel and kerosene consist of several hundred chemical species, estimating their heat transfer coefficient is virtually impossible unless simpler mixtures, known as surrogates, are used to emulate their targeted behavior. Following this strategy,we present in this paper a predictive method to estimate the nucleate boiling heat transfer coefficient with gasoline. In order to formulate surrogates, probability distribution functions were discretized to ground the choice of components and their respective molar fractions. At first, a traditional gamma distribution function was used but it provided unsatisfying results, so we developed an original one named double-log-normal distribution. Heat transfer coefficients and bubble points estimated at different pressures were compared with experimental data. The best surrogate was a 6-component mixture formulated with the proposed distribution and its application resulted in a 2.3% overall absolute deviation for the heat transfer coefficient estimation at 102 kPa. The model was validated at pressures up to 403 kPa. Finally, a sensitivity analysis of the double-log normal distribution showed the estimated nucleate boiling heat transfer coefficient is nearly the same regardless of gasoline composition variations.

... 30,31 Other research showed that a later injection timing would lead to the combustion deterio-ration and the rise of CO and HC emissions, but a small EDI ratio and EDI heating technology would solve these problems. 32,33 The above studies indicated that the application of ethanol in SI engines could improve the power and emission characteristics. However, fuel ethanol is usually made by fermentation, and the process of obtaining anhydrous ethanol from the fermentation broth consumes a lot of energy, which increases the production cost. ...

Decheng Li
Xiumin Yu
Ping Sun
Zhe Zhao

Ethanol is usually combined with gasoline to manufacture ethanol–gasoline with excellent combustion characteristics. However, extracting water from hydrous ethanol to manufacture anhydrous ethanol consumed much energy, which increases the production cost of ethanol–gasoline. Many researchers have studied the combustion and emissions of hydrous ethanol–gasoline to explore the application of hydrous ethanol–gasoline as the fuel for spark-ignition engines. Most previous studies changed the hydrous ethanol ratio with fixed purity in hydrous ethanol–gasoline to study the effects of hydrous ethanol. Different from previous studies, this paper studied the effects of water ratio (Wr) in hydrous ethanol on the combustion and emissions of a hydrous ethanol/gasoline combined injection engine under different excess air ratio (λ) values. The ratios of ethanol and gasoline keep constant, while the purity of hydrous ethanol changes during the research. The experiment adopted the combined injection mode with hydrous ethanol direct injection plus gasoline port injection; the direct injection ratio was 20%. The experiment set three λ (0.9, 1, and 1.2) and five Wrs (0, 5, 10, 15, and 20%). The test engine's speed was 1500 rpm, and the intake manifold absolute pressure was 48 kPa. Results showed that water inhibited combustion, prolonged CA 0-10 and CA 10-90, reduced Pmax and Tmax, and delayed APmax; larger λ made the deterioration on combustion more obvious, and the smaller λ had a larger tolerance to water. Water could increase torque and improve emissions, but different parameters corresponded to different optimal Wrs. For torque, the optimal Wr was 5%. For HC emissions, the optimal Wr was 0%; for CO emissions, the optimal value was 5%; and for NOx emissions, the best value was 20%. The best Wr was 10% for particle number (PN) emissions. Under the optimal Wr condition, when λ values were 0.9, 1, and 1.2, compared with pure gasoline, the torque increased by 7.5, 5.54, and 5.31%; HC emissions decreased by 21.37, 23.43, and 26.58%; NOx emissions decreased by 4.26, 11.47, and 12.55%; CO emissions decreased by 17.51, 34.56, −50%; and the total PN emissions decreased by 87.64, 89.64, and 76.07%.

... On the other hand, an explanation for the anomalous parameters in cluster 5 appears to be engine overcooling, in which the engine's normal operating temperature cannot be reached. Since engine overcooling can be just as damaging as engine overheating, this area warrants further investigation [53]. Additionally, based on the range of Maximum Continuity Rating (MCR) values converted from vessel engine rotation to percentage of power output, the vessel's condition was "slow ahead" with a risk of failure during deceleration and acceleration especially when the engine was started. ...

In this study, we proposed a data-driven approach to the condition monitoring of the marine engine. Although several unsupervised methods in the maritime industry have existed, the common limitation was the interpretation of the anomaly; they do not explain why the model classifies specific data instances as an anomaly. This study combines explainable AI techniques with anomaly detection algorithm to overcome the limitation above. As an explainable AI method, this study adopts Shapley Additive exPlanations (SHAP), which is theoretically solid and compatible with any kind of machine learning algorithm. SHAP enables us to measure the marginal contribution of each sensor variable to an anomaly. Thus, one can easily specify which sensor is responsible for the specific anomaly. To illustrate our framework, the actual sensor stream obtained from the cargo vessel collected over 10 months was analyzed. In this analysis, we performed hierarchical clustering analysis with transformed SHAP values to interpret and group common anomaly patterns. We showed that anomaly interpretation and segmentation using SHAP value provides more useful interpretation compared to the case without using SHAP value.

Jihwan Lee
Gian Antariksa

... Kabasin et al. [17] used an integrated heated fuel injector for ethanol direct injection engines, and it was found that flash boiling effects notably improved spray breakup and cold-start performance. Similar observations were established by Huang et al. combining port injection and direction injection schemes [18]. Fedor et al. [19] adopted a heated injector strategy combined with multiple injection schemes and found improvements in cold-start combustion characteristics. ...

Flash boiling atomization has the potential to enable superior spray atomization and more homogenous fuel-air mixing. These capabilities have been demonstrated in practical combustors such as gasoline direct-injection (GDI) engines. However, optimal injection schemes for flash boiling atomization has not been thoroughly investigated yet in the existing literature. In this work, we examined such contents in the aspects of split injection schemes coupled with flash boiling atomization. The impacts were examined firstly by velocity field analysis on single injection flash boiling plumes using particle imaging velocimetry (PIV). Furthermore, an optical reciprocating engine facility was used to study flame characteristics. Spray morphologies in the cylinder were studied, and the velocity field was visualized by the optical flow method. Important combustion characteristics, such as the flame speed and combustion duration, were investigated. Finally, the thermal efficiency under different injection schemes, as well as combustion emissions were quantified and compared to discuss the influence of flash boiling atomization on lean-operated GDI engines.

... I nternal combustion engines (ICEs) technology is progressing rapidly due to the more and more stringent [1,2,3,4] emission legislation. The introduction of WLTP and RDE vehicle emission test cycles represents an extreme challenge in engine design and calibration. ...

The new real driving emission cycles and the growing adoption of turbocharged GDI engines are directing the automotive technology towards the use of innovative solutions aimed at reducing environmental impact and increasing engine efficiency. Water injection is a solution that has received particular attention in recent years, because it allows to achieve fuel savings while meeting the most stringent emissions regulations. Water is able to reduce the temperature of the gases inside the cylinder, coupled with the beneficial effect of preventing knock occurrences. Moreover, water dilutes combustion, and varies the specific heat ratio of the working fluid; this allows the use of higher compression ratios, with more advanced and optimal spark timing, as well as eliminating the need of fuel enrichment at high load. Computational fluid dynamics simulations are a powerful tool to provide more in-depth details on the thermo-fluid dynamics involved in engine operations with water injection. The main intent of this work is to explore the effectiveness of port water injection installation in an optical access GDI engine operated with fully open throttle and fixed spark timing. Commercial gasoline is used as a fuel, while water is delivered through 2 injectors in the intake manifold. The injected water mass is 30% of the fuel mass. Numerical combustions are validated against experimental data, taking into account the blow-by effect with a crevices model, not negligible in an optical access engine. The G-equation turbulent combustion model is used in a RANS framework. Multi-cycle simulations are performed with water injection, also focusing on the wall film dynamics and the spray evolution. This study highlights the positive impact of water injection on lowering charge temperatures before ignition, with a consequent reduction in the peak pressure. Predicted heat release rates match measured data for both the baseline and the water injection cases. In addition, the CFD model allows to have additional insight on the in-cylinder processes. Remarkably, the analysis of the predicted flame front details agrees with the experimental imaging results in detecting an increase of the flame wrinkling and a transition in the combustion regime in the presence of water, despite the concurrent reduction of the overall burning rate.

... To address these issues, EDI heating was proposed as an economic and effective method to generate fine and fast evaporating DI sprays in the GPI+EDI engine [155]. The experimental results showed that ethanol fuel heating was an effective method to solve the problems of ethanol's slow evaporation and over-cooling effect in the EDI+GPI engine in terms of minimising the emissions. ...

Modern spark ignition engines mostly use one injection system to deliver gasoline into the combustion chamber, using either direct injection or port fuel injection. Both technologies have their respective advantages. To integrate their advantages and to promote the use of renewable fuels, dual injection engines are in development in recent years. Dual injection represents an advanced combustion system and is a novel technology to address the urgent issues of sustainability and environmental protection. This study reviews the state-of-the-art research on dual injection spark ignition engines with a focus on renewable fuels, their advantages and engine performance. The main advantages of dual injection include greater control flexibility, enhanced cooling effect, knock mitigation, engine downsizing, extended lean-burn limits, higher thermal efficiency and reductions of several emission species. The most promising renewable fuels for dual injection are ethanol, methanol and hydrogen. Each renewable fuel is aimed at different advantages of dual injection. Alcohol-gasoline dual injection provides great anti-knock ability by taking advantage of alcohols' large enthalpies of vaporisation and high octane numbers, while hydrogen-gasoline dual injection provides extended lean-burn limits by taking advantage of hydrogen's low ignition energy, wide flammability limit and high flame speed. Direct injection of renewable fuels is the optimal injection strategy because it effectively utilises the strong cooling effect of alcohols or avoids the volumetric efficiency reduction and pre-ignition of hydrogen. Dual injection generally demonstrates higher thermal efficiency than single injection. In addition, dual injection effectively reduces particulate emissions while there are usually trade-offs among gaseous emissions.

... In the treatment process to maximize vehicle performance by using of ethanol fuel there is one treatment that can be used without changing the component or system on vehicle that is heating the intake air in the combustion chamber, air intake heating aim to give energy on the air that will come into the combustion chamber [7] so that the intake air in combustion chamber has hot temperature because ethanol fuel has lower temperature than gasoline, the way of air intake works almost same with turbo charge process, but on this research the heating uses a nickelin wire that wrapped around air intake pipe, so when the air mixing with ethanol in the combustion chamber make the ethanol fuel temperature increasing by the mixture of hot air in the combustion chamber [3]. ...

Decheng Li
Xiumin Yu
Yaodong Du
Zhe Zhao

Hydrous ethanol has lower cost and energy consumption during the production process than anhydrous ethanol. Many researchers used hydrous ethanol instead of anhydrous ethanol to manufacture hydrous ethanol-gasoline. The water ratio in hydrous ethanol (ω) determines the energy consumption in the production process of ethanol. However, no one researched the effects of different ω on combustion and emissions of hydrous ethanol-gasoline. So, we prepared five kinds of ethanol with different ω (0, 5%, 10%, 15%, 20% vol.) to research the effects of ω on combustion and emissions. This study adopted combined injection with hydrous ethanol direct injection plus gasoline port injection (HEDI + GPI). The engine speed was 1500 rpm, and λ was 1, the intake manifold absolute pressure was 48 kPa, the direct injection ratio was 20%, five spark timings varied from 5 °CA BTDC to 25 °CA BTDC. The results showed hydrous ethanol prolonged flame development and propagation duration, decreased Tmax and Pmax, delayed APmax. Water increased COVimep, but using hydrous ethanol at the reasonable spark timing would not significantly affect the stability. Meanwhile, when the spark timing was MBT, compared with GE (gasoline/anhydrous ethanol dual fuel) fuel, as ω in GEW (gasoline/hydrous ethanol dual fuel) fuels increased, torque and brake thermal efficiency improved by 1.69%, 1.13%, −0.01%, −0.77%; HC increased by 2.13%, 9.02%, 23.89%, 37.15%; CO decreased by 41.95%, 28.56%, 5.59%, 2.46%; NOx decreased by 10.10%, −0.75%, −0.17%, 4.18%; total PN emissions decreased by 24.07%, 69.57%, 36.34%, 32.45%. Meanwhile, more water caused the size of particles to drop, made particles smaller.

Ze Liu
Ping Sun
Yaodong Du
Jiangdong Zhou

In this paper, a test platform for ethanol port injection plus gasoline direct injection was built to explore the effects of different ethanol-gasoline ratios (Re), direct injection timing (DIT) and ignition timing (IT) on the combustion and emissions of a SI engine. Results clearly show that when the total fuel contains more ethanol, the ignition timing corresponding to the maximum torque is smaller than the ignition timing when the total fuel contains more gasoline. Compared with the sole ethanol port injection (EPI) mode and gasoline direct injection (GDI) mode, G25 with EPI + GDI is the best mode for high-efficiency combustion and G25 with IT = 20°CA BTDC and DIT = 120°CA BTDC can be regarded as the optimal operating condition for highest torque output, braking thermal efficiency and lowest BSFC. As for exhaust emissions, CO and HC reach the lowest at G25 while NOx has an opposite trend. Delayed ignition within the range of 10-30°CA BTDC reduces HC and NOx emissions. As for particle number emissions, the size of peak particle number gradually decreases from G100 to G0, the accumulation particle number (APN) is further oxidized and shows a significant downward trend with the increase of ethanol. The lowest TPN was obtained at G25 in which the TPN has dropped by 9% and 85% respectively compared with sole EPI and GDI mode. When Re ≥ 50%, the nucleation mode particle number (NPN) in the total particle distribution is higher than the accumulation mode particle number, while there is an opposite result when Re＜50%. Overall, the findings of this paper make contributions to the development of high-efficiency and low-pollution combined engines and further promote the application of clean alternative fuels.

A one-dimensional model of a solenoid-driven common-rail diesel injector has been developed in order to study the influence of fuel temperature on the injection process. The model has been implemented after a thorough characterization of the injector, both from the dimensional and the hydraulic point of view. In this sense, experimental tools for the determination of the geometry of the injector lines and orifices have been described in the paper, together with the hydraulic setup introduced to characterize the flow behaviour through the calibrated orifices.

An experimental study is conducted in this paper in order to assess the influence of the fuel temperature on the performance of a last generation common-rail ballistic solenoid injector. Mass flow rate measurements are performed for a wide range of temperatures, extending from 253 to 373 K, representative of all the possible operating conditions of the injector in a real diesel engine, including cold start. The high pressure line and the injector holder were refrigerated, making it possible to carefully control the fuel temperature, whereas measurements at cold conditions were carried out with the help of a climatic chamber. Relevant features such as stationary mass flow, injection delay or the behaviour at the opening and closing stages are analysed together with parameters governing the flow, such as the injector discharge coefficient.

Masaru KUBOTA
Koji YOSHIDA
Hideo SHOJI
Hidenori Tanaka

The purpose of this study is to improve the engine performance and the exhaust gas emission characteristics of the diesel engine. The heated fuels are provided for the diesel engine in order to activate the fuel. The test engine is a four-stroke, direct fuel injection, diesel engine. The experiment is made under constant engine speed and const fuel injection volume. For an arbitrary mass flow rate of fuel injection, the fuel temperature is increased from 373 [K] to 673 [K] at 50 [K] intervals. The crank angles at ignition and maximum combustion pressure are delayed and the maximum combustion pressure is decreased as the fuel temperature rises. In cases of middle and large mass flow rate of fuel injection, the burning period is reduced and the brake thermal efficiency is decreased when the fuel temperature is higher than 573 [K]. However, NO_x concentration is gradually decreased as fuel temperature increases.

Concerning the throttling loss under part load conditions, it is feasible to further improve the engine thermal efficiency through operating the engine under the unthrottled condition and controlling its load by changing the excess air ratio. However, the narrow flammability of ethanol may lead the ethanol engine to encounter high cyclic variations under unthrottled and lean conditions. The addition of hydrogen is potentially helpful for solving this problem. In this test, the engine was run under an speed of 1400 rpm and unthrottled conditions. The hydrogen volume fractions in the intake were respectively kept at 0% and 3%. For a given hydrogen blending level, the ethanol flow rate was reduced to enable the engine to run under lean conditions. The results showed that the engine efficiency was improved with the blending of hydrogen. The highest thermal efficiency was improved by 6.07% after blending 3% hydrogen to the intake air. The addition of hydrogen could increase the engine torque output at lean conditions. Both cooling and exhaust losses were decreased after the hydrogen enrichment while adopting the lean combustion strategy. The hydrogen addition contributed to the extended lean burn limit and decreased cyclic variation under lean conditions. HC and CO emissions were decreased whereas NOx emissions were increased after the blending of hydrogen.

Ethanol direct injection (EDI) is a promising technology to address the issue of knock in downsized spark ignition (SI) engines due to the strong cooling effect of EDI and ethanol's large octane number. However, the evaporation rate of ethanol is lower than that of gasoline fuel because of its low volatility (saturation vapour pressure) in low temperature conditions and large enthalpy of vaporization. This might have caused the increased HC and CO emissions in an ethanol direct injection plus gasoline port injection (EDI+GPI) engine when EDI was applied. To address this issue, the combustion and emission performance of an EDI+GPI engine fuelled with hot ethanol fuel was experimentally investigated in the present study. The experiments were conducted on a 249 cc single cylinder SI engine at medium load (IMEP 6.0-6.3 bar) and stoichiometric fuel/air ratio condition. The injected ethanol fuel temperature ranged from 45 ℃ (no fuel heating) to 105 ℃ (flash-boiling spray) with an increment of 15 ℃. Experimental results showed that the IMEP decreased slightly with the increase of ethanol fuel temperature. However, the ISCO and ISHC emissions decreased significantly and ISNO increased moderately with the increase ethanol fuel temperature.

Roberto Krenus
Marcos R. V. Passos
Thiago Ortega
Kwang Han

After the second worldwide oil crisis, Brazil put in place by 1975 a strategic plan to stimulate the usage of ethanol (from sugar cane), to be mixed to the gasoline or to be sold as 100% ethanol fuel (known as E100). To enable an engine to operate with both gasoline and ethanol (and their mixtures), by 2003 the "Flex Fuel" technology was implemented. By 2012 calendar year, from a total of about 3.8 million vehicles sold in the Brazilian market, 91% offered the "Flex Fuel" technology, and great majority used a gasoline sub-tank to assist on cold starts (typically below 15°C, where more than 85% of ethanol is present in fuel tank). The gasoline sub-tank system suffers from issues such as gasoline deterioration, crash-worthiness and user inconvenience such as bad drivability during engine warm up phase. This paper presents fuel injector technologies capable of rapidly electrically heating the ethanol fuel for the Brazilian transportation market. These heated fuel injectors can be used for cold starting ethanol fueled engines as presented in SAE paper 2009-01-0615 [1] and to enable emissions reduction with a variety of automotive fuels as presented in SAE paper 2010-01-1265 [2]. This paper will demonstrate the benefits and advantages obtained by introducing the individually-controlled heated injectors on Hyundai HB20 1.6L Flex Fuel vehicles, which besides assisting on cold starts, also helps to meet upcoming emissions legal requirements (Proconve L6). Additionally, it will enable meeting future stringent regulations that will require unburned ethanol (ETOH) emission to be considered by Brazilian legislation. Through significant fuel injection optimization on both cold start and emissions, these advantages will also allow lowering fuel consumption with E100 and potential catalyst converter savings.

Source: https://www.researchgate.net/publication/304350043_Investigation_of_the_effect_of_heated_ethanol_fuel_on_combustion_and_emissions_of_an_ethanol_direct_injection_plus_gasoline_port_injection_EDI_GPI_engine

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