(3)

**2.4 WPPO properties analysis**

*List of equipment used in the experiment.*

The iodine value is according to Eq. (3):

*MW*i is the molecular weight.

**2.5 Engine testing and performance analysis**

The cetane index number is according to Eq. (4):

*A*i is the weight percentage of each fatty acid component. *D* is the number of double bonds in each fatty acid.

software. Combustion data was obtained, and graphs sketched.

is according to Eq. (2):

**Table 5.**

Where:

2000 r/min.

**88**

*Schematic diagram of the engine testing and equipment.*


#### **Table 6.**

*Experimental engine specifications.*

equivalent to 50% engine load. For intermediate speeds two speeds are chosen as 1500 r/min and full load at 2000 r/min for Mode 2 as 75% and 100% engine load equivalents respectively. For engine load the dynamometer is fitted with a screw type loading device enabling each load to be synchronized with the intended engine speed targeted. **Figure 2** shows the schematic of the test engine and **Table 6** shows the engine specification.

#### **3. Results and discussion**

#### **3.1 Brake specific fuel consumption (BSFC)**

Fuel consumption as BSFC is a measure of fuel flow per unit time measured as a flow rate. The fuel thus measures how an engine utilizes supplied fuel to produce the intended work. While measuring BSFC lower values are preferred compared to higher values. The brake specific fuel consumption measures the efficiency of fuel by combustion of the fuel and air mixture, which does the actual work of crankshaft rotation. In other word, the BSFC is a ratio of the rate of fuel consumption in relation to the effective power produced by the engine. This means for every cycle of operation the BSFC tries to get an equal output with the corresponding increase in fuel supply to the engine (the engine is supplied volumetrically).

**Figure 3** is a variation of BSFC with engine speed, and shows that as the speed increases, there is an equal increase in the fuel consumed by the test engine. The values obtained at full engine load for the blends of 95/WPPO5, 90/WPPO10, and 80/WPPO20, 70/WPPO30, 60/WPPO40 and PD were 0.4 kg/kW.h, 0.41 kg/kW.h, 0.42 kg/kW.h, 0.43 kg/kW.h and 0.35 kg/kW.h respectively. At high engine loads the conversion of heat energy to mechanical energy increased with increase in combustion temperature, leading to increased BSFC for the biodiesel. This increase was proportional to the difference in their heating values which is identical to the findings of [50]. Furthermore, the WPPO blends had high densities, therefore suffered high mass injection pressure, hence an increase in BSFC which is identical to studies by [51, 52]. WPPO blends compared well to conventional diesel fuel and sometimes other biodiesel blends with comparative differences in the heating values.

As the blend ratio increased there was a decrease in the BSFC across all the test fuels. However, the values for all WPPO blends increased compared to PD test fuel. This is due to the lower calorific values of the blends as the percentage of the blend ratio increased. In other words, by increasing the ratio of WPPO in the diesel test fuel the engine fuel consumption increased. This is identical to the studies of [53–55]. The closeness of the values and the packed graph reveal a close resemblance and identical BSFC characteristics of WPPO to PD properties. For example, **Figure 3** at Mode 1 (500 r/min (25% load) to 1000 r/min 50% engine load) blend 90/WPPO10 had a value of 0.48 kg/kW.h and 0.43 kg/kW.h compared to full engine speed Mode 2 2000 r/min

**91**

**Figure 4.**

*Brake thermal efficiency versus speed.*

*Assessing the Effects of Engine Load on Compression Ignition Engines Using Biodiesel Blends*

100% load) with 0.37 kg/kW.h and 0.41 kg/kW.h. This value is higher than PD with 0.4 kg/kW.h at 1000 r/min 50% engine load and 0.35 kg/kW.h at full engine load.

supplied by the fuel consumed by the engine. The brake thermal efficiency also determines how well the engine converts the heat energy into actual mechanical energy. The BTE is influenced by engine design, type of fuel used and the engine application [56]. High engine load seems to increase BTE as can be seen in intermediate loads of 1000 r/min 50% engine load to 1500 r/min 75% engine load. When operating at part load the gross thermal efficiency of any engine falls to 28%, which translates to a rating of 22% down, from a full load thermal efficiency at 39.1%. Modern on-road diesel engines provide a 42% BTE at full load but waste almost

The BTE measures the ratio of the engine brake power to the heat of combustion

The BTE variations with engine load is shown in **Figure 4**. The graphs show that as the load increased there was an increase in the BTE across all the test fuel blends of WPPO and PD. The result of this experiment shows that the BTE increased as the load increased, explained by the reduction in the heat loss as the engine power

At Mode 1 (1000 r/min, 50% engine load) the values for blends 95/WPPO5, 90/WPPO10, 80/WPPO20, 70/WPPO30, 60/WPPO40 and PD were 22%, 21%, 20%, 18%, 16.5% and 22.5% respectively. As the blend ratio and engine idling load increased there was an increase in BTE across the blends of WPPO, but with a decrease in the BTE within the blends. For example, at Mode 1 (500 r/min, 25% engine load), 95/WPPO5 had values of 14%, 22%, 26.5% and 25.7% compared to

This decrease in BTE within the blends is due to the presence of aromatic compounds in waste pyrolysis plastic oil, which require a lot of energy to break [18]. Another critical factor that could be contributing to lower BTE among blends of WPPO compared to PD fuel is the higher combustion temperature characteristics observed in WPPO fuel blends leading to high heat transfer losses [57]. The main factors causing reduction in the BTE with use of blends is their lower heating values, low air to fuel mixing (poor atomization of blends during injection), high

80/WPPO20 with 12.5%, 20%, 22.5% and 23% respectively.

viscosity, high biodiesel density, or higher BSFC [58].

*DOI: http://dx.doi.org/10.5772/intechopen.95974*

**3.2 Brake thermal efficiency (BTE)**

28% of all fuel used through exhaust gases.

(more fuel) increased with load.

**Figure 3.** *Brake specific fuel consumption versus speed.*

100% load) with 0.37 kg/kW.h and 0.41 kg/kW.h. This value is higher than PD with 0.4 kg/kW.h at 1000 r/min 50% engine load and 0.35 kg/kW.h at full engine load.

#### **3.2 Brake thermal efficiency (BTE)**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

in fuel supply to the engine (the engine is supplied volumetrically).

other biodiesel blends with comparative differences in the heating values.

As the blend ratio increased there was a decrease in the BSFC across all the test fuels. However, the values for all WPPO blends increased compared to PD test fuel. This is due to the lower calorific values of the blends as the percentage of the blend ratio increased. In other words, by increasing the ratio of WPPO in the diesel test fuel the engine fuel consumption increased. This is identical to the studies of [53–55]. The closeness of the values and the packed graph reveal a close resemblance and identical BSFC characteristics of WPPO to PD properties. For example, **Figure 3** at Mode 1 (500 r/min (25% load) to 1000 r/min 50% engine load) blend 90/WPPO10 had a value of 0.48 kg/kW.h and 0.43 kg/kW.h compared to full engine speed Mode 2 2000 r/min

Fuel consumption as BSFC is a measure of fuel flow per unit time measured as a flow rate. The fuel thus measures how an engine utilizes supplied fuel to produce the intended work. While measuring BSFC lower values are preferred compared to higher values. The brake specific fuel consumption measures the efficiency of fuel by combustion of the fuel and air mixture, which does the actual work of crankshaft rotation. In other word, the BSFC is a ratio of the rate of fuel consumption in relation to the effective power produced by the engine. This means for every cycle of operation the BSFC tries to get an equal output with the corresponding increase

**Figure 3** is a variation of BSFC with engine speed, and shows that as the speed increases, there is an equal increase in the fuel consumed by the test engine. The values obtained at full engine load for the blends of 95/WPPO5, 90/WPPO10, and 80/WPPO20, 70/WPPO30, 60/WPPO40 and PD were 0.4 kg/kW.h, 0.41 kg/kW.h, 0.42 kg/kW.h, 0.43 kg/kW.h and 0.35 kg/kW.h respectively. At high engine loads the conversion of heat energy to mechanical energy increased with increase in combustion temperature, leading to increased BSFC for the biodiesel. This increase was proportional to the difference in their heating values which is identical to the findings of [50]. Furthermore, the WPPO blends had high densities, therefore suffered high mass injection pressure, hence an increase in BSFC which is identical to studies by [51, 52]. WPPO blends compared well to conventional diesel fuel and sometimes

**3. Results and discussion**

**3.1 Brake specific fuel consumption (BSFC)**

**90**

**Figure 3.**

*Brake specific fuel consumption versus speed.*

The BTE measures the ratio of the engine brake power to the heat of combustion supplied by the fuel consumed by the engine. The brake thermal efficiency also determines how well the engine converts the heat energy into actual mechanical energy. The BTE is influenced by engine design, type of fuel used and the engine application [56]. High engine load seems to increase BTE as can be seen in intermediate loads of 1000 r/min 50% engine load to 1500 r/min 75% engine load. When operating at part load the gross thermal efficiency of any engine falls to 28%, which translates to a rating of 22% down, from a full load thermal efficiency at 39.1%. Modern on-road diesel engines provide a 42% BTE at full load but waste almost 28% of all fuel used through exhaust gases.

The BTE variations with engine load is shown in **Figure 4**. The graphs show that as the load increased there was an increase in the BTE across all the test fuel blends of WPPO and PD. The result of this experiment shows that the BTE increased as the load increased, explained by the reduction in the heat loss as the engine power (more fuel) increased with load.

At Mode 1 (1000 r/min, 50% engine load) the values for blends 95/WPPO5, 90/WPPO10, 80/WPPO20, 70/WPPO30, 60/WPPO40 and PD were 22%, 21%, 20%, 18%, 16.5% and 22.5% respectively. As the blend ratio and engine idling load increased there was an increase in BTE across the blends of WPPO, but with a decrease in the BTE within the blends. For example, at Mode 1 (500 r/min, 25% engine load), 95/WPPO5 had values of 14%, 22%, 26.5% and 25.7% compared to 80/WPPO20 with 12.5%, 20%, 22.5% and 23% respectively.

This decrease in BTE within the blends is due to the presence of aromatic compounds in waste pyrolysis plastic oil, which require a lot of energy to break [18]. Another critical factor that could be contributing to lower BTE among blends of WPPO compared to PD fuel is the higher combustion temperature characteristics observed in WPPO fuel blends leading to high heat transfer losses [57]. The main factors causing reduction in the BTE with use of blends is their lower heating values, low air to fuel mixing (poor atomization of blends during injection), high viscosity, high biodiesel density, or higher BSFC [58].

**Figure 4.** *Brake thermal efficiency versus speed.*

The highest BTE value was 24. 5% by blend 95/WPPO5 at 1500 r/min (Mode 2, 75% engine load) compared to any other blend of WPPO. **Figure 4** shows values of 24.8%, 23%, 21% and 19% for full speed 2000 r/min 100% engine load, Mode 2) respectively for blends 90/WPPO10, 80/WPPO20, 70/WPPO30, and 60/WPPO40. However, blend 60/WPPO40/E25 reported the lowest values compared to the other blends. At 500 r/min (Mode 1, 25% engine load), the BTE value was 9.5% compared to 19% at full load 2000 r/min, Mode 2). These two are the lowest values of BTE for all the blends tested, as shown in **Figure 4**.

#### **3.3 Carbon monoxide (CO)**

**Figure 5** is a variation of CO with two engine load modes (Mode 1, and Mode 2) with speed range of 500 r/min 25% engine load to 2000 r/min 100% engine load. The graph reveals that as the engine speed, load and the blend ratio increased, CO emissions reduced up to engine speeds of 1500 r/min (Mode 2, 75% of engine load). This was for PD and all blends 95/WPPO5, 90/WPPO10, 80/WPPO20, 70/WPPO30, and 60/WPPO40; the values were 270 ppm, 285 ppm, 315 ppm, 345 ppm, 370 ppm, 385 ppm respectively. The highest value of CO emission reported was 485 ppm for blend 60/WPPO40 and the lowest value reported was for blend 95/WPPO5 at 388 ppm.

Another observation is that as the engine was approaching full load (Mode 2, 2000 r/min), all the test fuels showed increased CO emissions with blends 95/WPPO5 and 90/WPPO10 reporting the lowest emissions value of 388 ppm and 435 ppm among the test blends across the entire engine load Modes 1 and 2 conditions. However, as the load increased from Mode 1 25% engine load speed of 500 r/min) to Mode 2 (75% engine load) the values reported were 320 ppm and 335 ppm respectively.

There are a number of factors, which explain the low CO emissions as the engine load and speed is increasing. The reason the blends show decreasing and increasing trends for Modes 1 and 2 respectively is due to high viscosity in WPPO. Viscosity affects the spray pattern resulting in poor fuel mixing therefore incomplete combustion and increased carbon monoxide emissions [59]. This phenomenon is linked to the increased engine load and the short ignition delay, hence increasing CO emissions. Additionally, the decrease in CO emissions could be due to the conversion of CO to CO2 taking up this reaction from the high oxygen content of biodiesel [60].

**93**

**Figure 6.**

*Unburnt hydrocarbons versus speed.*

*Assessing the Effects of Engine Load on Compression Ignition Engines Using Biodiesel Blends*

UHC exhaust emissions are due to poor atomization after injection, over leaning zones and wall flame quenching [61, 62]. **Figure 6** is a variation of UHC emission with engine load. As the engine load and engine speed increased, the UHC emissions increased too. However, the increase was more significant as the engine load was in intermediate loads Mode 1 and Mode 2, 1000 r/min to 1500 r/min full load (50–75%). For example, at Mode 1, (1000 r/min, 50% engine load), the blend values were 23 ppm, 24 ppm, 26 ppm, 28 ppm, and 30 ppm respectively compared to full load Mode 2 2000 r/min) with 13 ppm, 15 ppm, 16 ppm, 18 ppm, and 26 ppm for blends 95/WPPO5, 90/WPPO10, 80/WPPO20, 70/WPPO30, and 60/WPPO40 respectively. The blends 95/WPPO5 and 90/WPPO10 produced lower UHC emissions compared to the other test blends. The trends in **Figure 6** show high emission values for the blends compared to the PD test fuel values. However, the general trend in **Figure 6** shows that increased blend ratio significantly reduced UHC emissions across all the test fuels irrespective of the engine Mode. This reduction is due to the high oxygen of WPPO which has an oxygen content of 7.83 as shown in **Table 6** and

There are two main causes of increased hydrocarbon emissions, due to hydrogen radicals in diesel-WPPO blends and the presence of higher aromatic compounds [63]. Another contribution is caused by high density, low viscosity and low cetane of WPPO blends resulting in poor spray characteristics, leading to wall impingement, thus high UHC emissions. High blend ratio has also been identified as a factor that influences formation of UHC emissions using WPPO and has been reported by a number of researchers such as [64–66]. Hence the conclusion that high engine loads increases the values of UHC emissions proportionately to petroleum diesel. Increased UHC emissions can also be attributed to engine operating environment especially if the temperature range of 400–600°C exists in the combustion chamber. This is due to diesel exhaust pipe reaction, which either lowers or increases the

It is an established fact that NOX emission are a function of in-cylinder temperature and atmospheric nitrogen, which is at 78% during intake. NOx emissions are also a

*DOI: http://dx.doi.org/10.5772/intechopen.95974*

in Section 3 of the results and discussion.

concentration of UHC [67, 68].

**3.5 Oxides of nitrogen (NOX)**

**3.4 Unburnt hydrocarbons (UHC)**

**Figure 5.** *Carbon monoxide versus speed.*

*Assessing the Effects of Engine Load on Compression Ignition Engines Using Biodiesel Blends DOI: http://dx.doi.org/10.5772/intechopen.95974*

#### **3.4 Unburnt hydrocarbons (UHC)**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

all the blends tested, as shown in **Figure 4**.

**3.3 Carbon monoxide (CO)**

blend 95/WPPO5 at 388 ppm.

335 ppm respectively.

The highest BTE value was 24. 5% by blend 95/WPPO5 at 1500 r/min (Mode 2, 75% engine load) compared to any other blend of WPPO. **Figure 4** shows values of 24.8%, 23%, 21% and 19% for full speed 2000 r/min 100% engine load, Mode 2) respectively for blends 90/WPPO10, 80/WPPO20, 70/WPPO30, and 60/WPPO40. However, blend 60/WPPO40/E25 reported the lowest values compared to the other blends. At 500 r/min (Mode 1, 25% engine load), the BTE value was 9.5% compared to 19% at full load 2000 r/min, Mode 2). These two are the lowest values of BTE for

**Figure 5** is a variation of CO with two engine load modes (Mode 1, and Mode 2) with speed range of 500 r/min 25% engine load to 2000 r/min 100% engine load. The graph reveals that as the engine speed, load and the blend ratio increased, CO emissions reduced up to engine speeds of 1500 r/min (Mode 2, 75% of engine load). This was for PD and all blends 95/WPPO5, 90/WPPO10, 80/WPPO20, 70/WPPO30, and 60/WPPO40; the values were 270 ppm, 285 ppm, 315 ppm, 345 ppm, 370 ppm, 385 ppm respectively. The highest value of CO emission

reported was 485 ppm for blend 60/WPPO40 and the lowest value reported was for

Another observation is that as the engine was approaching full load (Mode 2, 2000 r/min), all the test fuels showed increased CO emissions with blends 95/WPPO5 and 90/WPPO10 reporting the lowest emissions value of 388 ppm and 435 ppm among the test blends across the entire engine load Modes 1 and 2 conditions. However, as the load increased from Mode 1 25% engine load speed of 500 r/min) to Mode 2 (75% engine load) the values reported were 320 ppm and

There are a number of factors, which explain the low CO emissions as the engine load and speed is increasing. The reason the blends show decreasing and increasing trends for Modes 1 and 2 respectively is due to high viscosity in WPPO. Viscosity affects the spray pattern resulting in poor fuel mixing therefore incomplete combustion and increased carbon monoxide emissions [59]. This phenomenon is linked to the increased engine load and the short ignition delay, hence increasing CO emissions. Additionally, the decrease in CO emissions could be due to the conversion of CO to CO2 taking up this reaction from the high oxygen content of biodiesel [60].

**92**

**Figure 5.**

*Carbon monoxide versus speed.*

UHC exhaust emissions are due to poor atomization after injection, over leaning zones and wall flame quenching [61, 62]. **Figure 6** is a variation of UHC emission with engine load. As the engine load and engine speed increased, the UHC emissions increased too. However, the increase was more significant as the engine load was in intermediate loads Mode 1 and Mode 2, 1000 r/min to 1500 r/min full load (50–75%). For example, at Mode 1, (1000 r/min, 50% engine load), the blend values were 23 ppm, 24 ppm, 26 ppm, 28 ppm, and 30 ppm respectively compared to full load Mode 2 2000 r/min) with 13 ppm, 15 ppm, 16 ppm, 18 ppm, and 26 ppm for blends 95/WPPO5, 90/WPPO10, 80/WPPO20, 70/WPPO30, and 60/WPPO40 respectively.

The blends 95/WPPO5 and 90/WPPO10 produced lower UHC emissions compared to the other test blends. The trends in **Figure 6** show high emission values for the blends compared to the PD test fuel values. However, the general trend in **Figure 6** shows that increased blend ratio significantly reduced UHC emissions across all the test fuels irrespective of the engine Mode. This reduction is due to the high oxygen of WPPO which has an oxygen content of 7.83 as shown in **Table 6** and in Section 3 of the results and discussion.

There are two main causes of increased hydrocarbon emissions, due to hydrogen radicals in diesel-WPPO blends and the presence of higher aromatic compounds [63]. Another contribution is caused by high density, low viscosity and low cetane of WPPO blends resulting in poor spray characteristics, leading to wall impingement, thus high UHC emissions. High blend ratio has also been identified as a factor that influences formation of UHC emissions using WPPO and has been reported by a number of researchers such as [64–66]. Hence the conclusion that high engine loads increases the values of UHC emissions proportionately to petroleum diesel. Increased UHC emissions can also be attributed to engine operating environment especially if the temperature range of 400–600°C exists in the combustion chamber. This is due to diesel exhaust pipe reaction, which either lowers or increases the concentration of UHC [67, 68].

#### **3.5 Oxides of nitrogen (NOX)**

It is an established fact that NOX emission are a function of in-cylinder temperature and atmospheric nitrogen, which is at 78% during intake. NOx emissions are also a

**Figure 6.** *Unburnt hydrocarbons versus speed.*

function of three main mechanisms in the combustion theory [61, 69]. **Figure 7** is a variation of engine idling load with NOX emissions. The figure shows that as the engine idling load was increased there was an increase in the NOX emissions irrespective of fuel blend ratio. The values of NOX emissions for the blends 95/WPPO5, 90/WPPO10, and 80/WPPO20 reported higher values at (Mode 2, 75% load) compared to Mode 1. For example, at 1500 the values of the blends were 335 ppm, 358 ppm, and 475 ppm, compared to PD fuel at 300 ppm.

**Table 7** is showing different test fuel properties, units of measurement and testing standards used in this experiment. Blends 70/WPPO30 and 60/WPPO40 had the highest NOX emissions compared to the other blends of 95/WPPO5, 90/WPPO10, and 80/WPPO20 across all the engine load conditions tested.



#### **Table 7.**

*Test fuel biodiesel properties, units of measurement, testing standard methods and the values for PD compared to WPPO.*

**95**

**Figure 8.**

*In-cylinder pressure vs. crank angle variation compared to diesel and WPPO.*

*Assessing the Effects of Engine Load on Compression Ignition Engines Using Biodiesel Blends*

At 500 r/min 25% engine load (Mode 1), the two blends (70/WPPO30, 60/WPPO40) had values of 235 ppm and 255 ppm respectively. However, at full speed (2000 r/min, 100% engine load (Mode 2) the NOX emissions for the two blends increased to 490 ppm and 525 ppm respectively compared to blends 95/WPPO5 for the same speed and engine load (500 r/min, Mode 1) at 175 ppm and at full load (2000 r/min,

As the blend ratio in **Figure 7** increased there was a direct increase in emissions of NOX across all the blended fuels. However, blend 95/WPPO5 and 90/WPPO10 reported the lowest values of NOX emissions (175 ppm and 195 ppm) compared to all the other tested blends experimented. The formation of NOX in biodiesel fuel combustion strongly depends on the combustion temperatures and the oxygen concentration in the combustion zone. The high blend ratios of 80/WPPO20, 70/WPPO30, and 60/WPPO40 showed a shortened combustion process. Thus, a poor cooling effect and failure to decrease peak combustion temperatures leads to increased NOX. WPPO blends emitted higher NOX due to the higher cetane index compared to diesel fuel. High cetane index number fuels have a shorter ignition delay which means longer residence time at elevated chamber temperatures, hence higher NOX compared to PD. The increased NOX emissions are a result of the presence of increased cetane index [70, 71] and other contaminants from the WPPO biodiesel impurities. Additionally, this could be due to the generation of radicals of hydrocarbons

through molecular unsaturation in the blends, this being identical to the findings of [72, 73]. The final factor is due to increased chamber temperature which improves combustion but increases NOX emissions, linked to the high oxygen content and the

Due to the high cetane index of WPPO biodiesel the combustion process starts early compared to PD, hence a higher release rate than PD combustion. This leads to a higher cylinder peak pressure for WPPO biodiesel fuel compared to PD. Depicted in **Figure 8**, is a comparison of s WPPO blends with PD in Mode 1 25% engine load at speeds 500 r/min to 1000 r/min. Under this condition, WPPO blends in Mode 1

*DOI: http://dx.doi.org/10.5772/intechopen.95974*

Mode 2) at 345 ppm.

air fuel ratio factors [49].

**4. WPPO combustion analysis**

*Assessing the Effects of Engine Load on Compression Ignition Engines Using Biodiesel Blends DOI: http://dx.doi.org/10.5772/intechopen.95974*

At 500 r/min 25% engine load (Mode 1), the two blends (70/WPPO30, 60/WPPO40) had values of 235 ppm and 255 ppm respectively. However, at full speed (2000 r/min, 100% engine load (Mode 2) the NOX emissions for the two blends increased to 490 ppm and 525 ppm respectively compared to blends 95/WPPO5 for the same speed and engine load (500 r/min, Mode 1) at 175 ppm and at full load (2000 r/min, Mode 2) at 345 ppm.

As the blend ratio in **Figure 7** increased there was a direct increase in emissions of NOX across all the blended fuels. However, blend 95/WPPO5 and 90/WPPO10 reported the lowest values of NOX emissions (175 ppm and 195 ppm) compared to all the other tested blends experimented. The formation of NOX in biodiesel fuel combustion strongly depends on the combustion temperatures and the oxygen concentration in the combustion zone. The high blend ratios of 80/WPPO20, 70/WPPO30, and 60/WPPO40 showed a shortened combustion process. Thus, a poor cooling effect and failure to decrease peak combustion temperatures leads to increased NOX. WPPO blends emitted higher NOX due to the higher cetane index compared to diesel fuel. High cetane index number fuels have a shorter ignition delay which means longer residence time at elevated chamber temperatures, hence higher NOX compared to PD.

The increased NOX emissions are a result of the presence of increased cetane index [70, 71] and other contaminants from the WPPO biodiesel impurities. Additionally, this could be due to the generation of radicals of hydrocarbons through molecular unsaturation in the blends, this being identical to the findings of [72, 73]. The final factor is due to increased chamber temperature which improves combustion but increases NOX emissions, linked to the high oxygen content and the air fuel ratio factors [49].

#### **4. WPPO combustion analysis**

Due to the high cetane index of WPPO biodiesel the combustion process starts early compared to PD, hence a higher release rate than PD combustion. This leads to a higher cylinder peak pressure for WPPO biodiesel fuel compared to PD. Depicted in **Figure 8**, is a comparison of s WPPO blends with PD in Mode 1 25% engine load at speeds 500 r/min to 1000 r/min. Under this condition, WPPO blends in Mode 1

**Figure 8.** *In-cylinder pressure vs. crank angle variation compared to diesel and WPPO.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

function of three main mechanisms in the combustion theory [61, 69]. **Figure 7** is a variation of engine idling load with NOX emissions. The figure shows that as the engine idling load was increased there was an increase in the NOX emissions irrespective of fuel blend ratio. The values of NOX emissions for the blends 95/WPPO5, 90/WPPO10, and 80/WPPO20 reported higher values at (Mode 2, 75% load) compared to Mode 1. For example, at 1500 the values of the blends were 335 ppm, 358 ppm, and 475 ppm,

**Table 7** is showing different test fuel properties, units of measurement and testing standards used in this experiment. Blends 70/WPPO30 and 60/WPPO40 had the highest NOX emissions compared to the other blends of 95/WPPO5, 90/WPPO10, and 80/WPPO20 across all the engine load conditions tested.

**Property Unit PD WPPO Standard** Appearance — Clear/brown Clear/amber Visual Density @ 20 °C kg/M3 838.8 788.9 ASTM D1298

Cetane index — 46 65a ASTM D4737 Hydrogen % 12.38 11.77 ASTM D7171 Cu corrosion 3 hrs @ 100°C — 1B ASTM D130 Carbon % 74.99 79.60 ASTM D7662 Oxygen % 12.45 7.83 ASTM D5622 Sulfur content % < 0.0124 0.15 ASTM D4294

Recovery % 98 — Residue and loss % 2.0 — Gross calorific value MJ/kg 44.84 42.15b ASTM D4868

*Test fuel biodiesel properties, units of measurement, testing standard methods and the values for PD compared* 

/s 2.32 2.17 ASTM D445

C 56.0 20.0 ASTM D93

C 160 119 ASTM D86

C 353.5 353.5 ASTM D86

Kinematic Visc @ 40 °C mm2

*Oxides of nitrogen emissions versus different engine speeds.*

compared to PD fuel at 300 ppm.

Flash point <sup>0</sup>

IBP temperature <sup>0</sup>

FBP temperature <sup>0</sup>

*a and b are calculated values.*

**94**

**Table 7.**

**Figure 7.**

*to WPPO.*

#### **Figure 9.**

*In-cylinder pressure vs. crank angle variation compared to diesel and WPPO in mode 2.*

exhibited higher peak cylinder pressure compared to PD, which is evident as the blend ratio increased as in **Figure 8**.

Compared to when the engine is running at high speed (high load), low speeds and low load residual gas temperatures and engine wall temperatures are low [69]. In other words, injection pressure and fuel temperature are low hence increased delay. This is the explanation why diesel in the combustion analysis starts after 3° CA compared to WPPO biodiesel blends. This causes diesel fuel to reach peak cylinder pressure after top dead center (TDC) in the power stroke. On the other hand, biodiesel blends reach peak cylinder pressure early, i.e., before TDC in the power stroke. For example, in **Figure 9** Mode 2 speed, the value for peak cylinder pressure for PD fuel is 55 bar compared to 56 bar for WPPO blend 95/WPPO5. This is due to enhanced combustion resulting from rapid combustion of the biodiesel blends at the pre-mixed phase. Of all the test fuels PD had the lowest peak cylinder pressure which occurs slightly after TDC [74].

#### **5. Conclusion**


**97**

**Author details**

emissions.

engine Mode.

South Africa

Semakula Maroa\* and Freddie Inambao

provided the original work is properly cited.

College of Agriculture Science and Engineering, Discipline of Mechanical Engineering, Green Energy Group, University of KwaZulu-Natal, Durban,

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium,

\*Address all correspondence to: ssemakulamaroa@gmail.com

*Assessing the Effects of Engine Load on Compression Ignition Engines Using Biodiesel Blends*

1000 r/min 50% engine load to 1500 r/min 75% engine load.

pyrolysis plastic oil, which require a lot of energy to break.

load and speeds seems to increase BTE as can be seen in intermediate loads of

• Result obtained during this experiment show that the BTE increased as the load increased. This is explained by the reduction in the heat loss as the engine

• As the blend ratio and engine load and speed increased there was an increase in BTE across the blends of WPPO, but with a decrease in the BTE within the blends. This was attributed to the presence of aromatic compounds in waste

• During experimentation it was observed that as the engine was approaching full load (Mode 2, 2000 r/min), all the test fuels showed increased CO

• As the engine load and engine speed increased, the UHC emissions increased too. However, the general trend in **Figure 6** shows that increased blend ratio significantly reduced UHC emissions across all the test fuels irrespective of the

• As the engine load and speed increased there was an increase in the NOX

• As the blend ratio increased there was a direct increase in emissions of NOX

*DOI: http://dx.doi.org/10.5772/intechopen.95974*

power (more fuel) increased with load.

emissions irrespective of fuel blend ratio.

across all the blended fuels.

load and speeds seems to increase BTE as can be seen in intermediate loads of 1000 r/min 50% engine load to 1500 r/min 75% engine load.


### **Author details**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

exhibited higher peak cylinder pressure compared to PD, which is evident as the

*In-cylinder pressure vs. crank angle variation compared to diesel and WPPO in mode 2.*

Compared to when the engine is running at high speed (high load), low speeds and low load residual gas temperatures and engine wall temperatures are low [69]. In other words, injection pressure and fuel temperature are low hence increased delay. This is the explanation why diesel in the combustion analysis starts after 3° CA compared to WPPO biodiesel blends. This causes diesel fuel to reach peak cylinder pressure after top dead center (TDC) in the power stroke. On the other hand, biodiesel blends reach peak cylinder pressure early, i.e., before TDC in the power stroke. For example, in **Figure 9** Mode 2 speed, the value for peak cylinder pressure for PD fuel is 55 bar compared to 56 bar for WPPO blend 95/WPPO5. This is due to enhanced combustion resulting from rapid combustion of the biodiesel blends at the pre-mixed phase. Of all the test fuels PD had the lowest peak cylinder

• In the discussions in Section 3 under BSFC it was observed that, at high engine loads the conversion of heat energy to mechanical energy increased with increase in combustion temperature, leading to increased BSFC for the biodiesel. The increase was found to be proportional to the heating values of the

• Due to high densities exhibited by most biodiesel blends it was observed that blends suffer from high mass injection pressures which is in return increases

increase in engine fuel consumption due to lower calorific values of the blends. However, the values for all WPPO blends increased compared to PD test fuel.

• The Brake thermal efficiency of diesel engines was observed to be influenced by engine design, type of fuel used and the engine application. High engine

• As the percentage of the blend ratio increased there was a proportionate

blend ratio increased as in **Figure 8**.

pressure which occurs slightly after TDC [74].

**5. Conclusion**

**Figure 9.**

different test fuels.

the BSFCs of blends.

**96**

Semakula Maroa\* and Freddie Inambao College of Agriculture Science and Engineering, Discipline of Mechanical Engineering, Green Energy Group, University of KwaZulu-Natal, Durban, South Africa

\*Address all correspondence to: ssemakulamaroa@gmail.com

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

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[49] Devan P, Mahalakshmi N. A Study of the Performance, Emission and Combustion Characteristics of a Compression Ignition Engine Using Methyl Ester of Paradise Oil–Eucalyptus Oil Blends. Applied Energy 2009;86(5) 675-680.

[50] Lapuerta M, Armas O, Hernández JJ, Tsolakis A. Potential for Reducing Emissions in a Diesel Engine by Fuelling with Conventional Biodiesel and Fischer–Tropsch Diesel. Fuel 2010;89(10) 3106-3113.

[51] Raheman H, Ghadge SV. Performance of Compression Ignition Engine with Mahua (*Madhuca indica*) Biodiesel. Fuel 2007;86(16) 2568-2573.

[52] Tüccar G, Tosun E, Özgür T, Aydın K. 2014 Diesel Engine Emissions and Performance from Blends of *Citrus sinensis* Biodiesel and Diesel Fuel. Fuel 2014;132 7-11.

[53] Armas O, Yehliu K, Boehman AL. Effect of Alternative Fuels on Exhaust Emissions During Diesel Engine Operation with Matched Combustion Phasing. Fuel 2010;89(2) 438-456.

[54] Godiganur S, Murthy CS, Reddy RP. Performance and Emission Characteristics of a Kirloskar HA394 Diesel Engine Operated on Fish Oil Methyl Esters. Renewable Energy 2010;5(2) 355-359.

[55] Zhu L, Zhang W, Liu W, Huang Z. 2010 Experimental Study on Particulate and NOx Emissions of a Diesel Engine Fueled with Ultra Low Sulfur Diesel, RME-Diesel Blends and PME-Diesel Blends. Science of the Total Environment 2010; 408(5) 1050-1058.

[56] Ramalingam S, Rajendran S. 2019 Assessment of Performance, Combustion, and Emission Behavior of Novel Annona Biodiesel-Operated Diesel Engine. In: Azad K (ed.) Advances in Eco-Fuels for a Sustainable Environment. Amsterdam: Elsevier; 2019. p391-405.

[57] Kalargaris I, Tian G, Gu S. Combustion, Performance and Emission Analysis of a DI Diesel Engine Using Plastic Pyrolysis Oil. Fuel Processing Technology 2017;157 108-115.

[58] Uddin SA, Azad AK, Alam M, Ahamed J. Performance of a Diesel Engine Run with Mustard-Kerosene Blends. Procedia Engineering 2015;105 698-704.

[59] Ghurri A, Kim J-D, Kim HG, Jung, J-Y, Song K-K. The Effect of Injection Pressure and Fuel Viscosity on the Spray Characteristics of Biodiesel Blends Injected into an Atmospheric Chamber. Journal of Mechanical Science and Technology 2012;26(9) 2941-2947.

[60] Rahman MM, Hassan MH, Kalam MA, Atabani AE, Memon LA, Rahman, SA. Performance and Emission Analysis of *Jatropha curcas* and *Moringa oleifera* Methyl Ester Fuel Blends in a Multi-Cylinder Diesel Engine. Journal of Cleaner Production 2014;65 304-310.

[61] Heywood JB. 2012 Internal Combustion Engine Fundamentals (Vol. 17). New Delhi, India: McGraw Hill Education India.

[62] Mendez S, Kashdan JT, Bruneaux G, Thirouard B, Vangraefschepe F. Formation of Unburned Hydrocarbons in Low Temperature Diesel Combustion. SAE International Journal of Engines 2010;2(2) 205-225.

[63] Murugan S, Ramaswamy M, Nagarajan G. The use of Tyre Pyrolysis Oil in Diesel Engines. Waste Management 2008;28(12) 2743-2749.

[64] Lujaji, F, Kristóf L, Bereczky A, Mbarawa M. Experimental Investigation of Fuel Properties, Engine Performance, Combustion and Emissions of Blends Containing Croton Oil, Butanol, and Diesel on a CI Engine. Fuel 2011;90(2) 505-510.

[65] Maroa S, Inambao F. Effects of Biodiesel Blends Varied by Cetane Numbers and Oxygen Contents on Stationary Diesel Engine Performance and Exhaust Emissions. In: Wos P, Jakubowski M (eds) Numerical and Experimental Studies on Combustion Engines and Vehicles. Rijeka: IntechOpen; 2020. DOI: 10.5772/ intechopen.92569

[66] Tutak W, Lukacs K, Szwaja S, Bereczky, A. Alcohol–Diesel Fuel Combustion in the Compression Ignition Engine. Fuel 2015;154 196-206.

[67] Sanli H, Canakci M, Alptekin E, Turkcan A, Ozsezen AN. Effects of Waste Frying Oil Based Methyl and Ethyl Ester Biodiesel Fuels on the Performance, Combustion and Emission Characteristics of a DI Diesel Engine. Fuel 2015;159 179-187.

[68] Shirneshan A. HC, CO, CO2 and NOx Emission Evaluation of a Diesel Engine Fueled with Waste Frying Oil Methyl Ester. Procedia-Social and Behavioral Sciences 2013;75 292-297.

[69] Turns SR. An Introduction to Combustion: Concepts and Applications (3 ed.) New Delhi, India: McGraw-Hill; 1996.

[70] Hoekman SK, Broch A, Robbins C, Ceniceros E, Natarajan M. Review of Biodiesel Composition, Properties, and Specifications. Renewable and Sustainable Energy Reviews 2012;16(1) 143-169.

[71] Robbins C, Hoekman SK, Gertler A, Broch A, Natarajan M. 2009 Biodistillate Transportation Fuels 2.-Emissions Impacts (No. 2009-01- 2724). SAE Technical Paper.

[72] Altun Ş. Effect of the Degree of Unsaturation of Biodiesel Fuels on the Exhaust Emissions of a Diesel Power Generator. Fuel 2014;117 450-457.

[73] Benjumea P, Agudelo JR, Agudelo AF. Effect of the Degree of Unsaturation of Biodiesel Fuels on Engine Performance, Combustion Characteristics, and Emissions. Energy & Fuels 2010;25(1) 77-85.

[74] Tarabet L, Loubar K, Lounici M, Hanchi S, Tazerou, M. Experimental evaluation of Performance and Emissions of DI Diesel Engine Fuelled with Eucalyptus Biodiesel. Proceedings of the Internal Combustion Engines: Performance, Fuel Economy and Emissions 2011;167-176.

**103**

**Chapter 6**

**Abstract**

*Jayashri N. Nair*

discussed followed by biodiesel.

tions on combustion process and emissions.

multiple injections

**1. Introduction**

Mitigation of Emissions through

Injection Strategies for C I Engine

Fuel conversion efficiency is high with diesel engines compared to petrol engines. However high emissions from diesel is a matter of concern and its mitigation paves way for scope of research. Exhaust gas recirculation is one of the method

widely accepted to curb NOx emissions. Recently, split or multiple-injection strategy has been explored by researchers to precisely control the fuel injected per cycle and also to mitigate emissions. Present work reflects technical review of effect of injection strategies on performance, emissions and combustion on C.I. engine with diesel and biodiesel as fuel. Injection strategies like duration of injection, number of injections, the dwell period between two injections, quantity of injection, and multiple injections are analyzed for their influence on engine output and brake specific fuel consumption. Also their effect on emissions especially soot and NOx emission are reviewed. First the effect of injection strategies with diesel fuel is

**Keywords:** split injection, pilot injection, NOx emissions, Soot emissions,

The emissions from diesel engine are hazardous for human health as it causes respiratory problem in human beings. NOx emissions in combination with water and oxygen forms acid rain and is also responsible for the global warming. To improve the air quality in cities, regulations and control measures are taken to lower exhaust emissions from heavy vehicles. However, there are no much improvements mostly because of the difficulties in eliminating NOx and PM emissions. The need of an hour is simultaneous reduction of NOx and PM emissions from diesel engines. However it is challenging to control NOx and PM emission simultaneously. Optimizing injection timing through split injection can be a prominent solution to curb these emissions from engine. In split injection, the injection is divided into two or more pulses thus reducing the delay time. Major fraction of the combustion takes place during expansion stroke. NOx released from the engine is reduced as NOx is mostly formed in the premixed combustion. Investigations are done by many researchers to optimize the injection scheme and analyze the effect of split injec-

Fuel injection timings influence the performance, emissions and the combustion part of the engine to a great extent. With the variation in the injection timings the state of air changes which effects the ignition delay. Occurrence of injection when the piston is far away from TDC reduces the temperature and pressure of the air

#### **Chapter 6**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

[69] Turns SR. An Introduction to Combustion: Concepts and Applications (3 ed.) New Delhi, India: McGraw-Hill;

[70] Hoekman SK, Broch A, Robbins C, Ceniceros E, Natarajan M. Review of Biodiesel Composition, Properties, and Specifications. Renewable and Sustainable Energy Reviews 2012;16(1)

[71] Robbins C, Hoekman SK, Gertler A, Broch A, Natarajan M. 2009 Biodistillate Transportation Fuels 2.-Emissions Impacts (No. 2009-01- 2724). SAE

[72] Altun Ş. Effect of the Degree of Unsaturation of Biodiesel Fuels on the Exhaust Emissions of a Diesel Power Generator. Fuel 2014;117 450-457.

[73] Benjumea P, Agudelo JR, Agudelo AF. Effect of the Degree of Unsaturation of Biodiesel Fuels on Engine Performance, Combustion Characteristics, and Emissions. Energy

& Fuels 2010;25(1) 77-85.

Emissions 2011;167-176.

[74] Tarabet L, Loubar K, Lounici M, Hanchi S, Tazerou, M. Experimental evaluation of Performance and Emissions of DI Diesel Engine Fuelled with Eucalyptus Biodiesel. Proceedings of the Internal Combustion Engines: Performance, Fuel Economy and

1996.

143-169.

Technical Paper.

(Vol. 17). New Delhi, India: McGraw

[63] Murugan S, Ramaswamy M, Nagarajan G. The use of Tyre Pyrolysis

Management 2008;28(12) 2743-2749.

[64] Lujaji, F, Kristóf L, Bereczky A, Mbarawa M. Experimental Investigation of Fuel Properties, Engine Performance, Combustion and Emissions of Blends Containing Croton Oil, Butanol, and Diesel on a CI Engine. Fuel 2011;90(2)

[65] Maroa S, Inambao F. Effects of Biodiesel Blends Varied by Cetane Numbers and Oxygen Contents on Stationary Diesel Engine Performance and Exhaust Emissions. In: Wos P, Jakubowski M (eds) Numerical and Experimental Studies on Combustion

Engines and Vehicles. Rijeka: IntechOpen; 2020. DOI: 10.5772/

[66] Tutak W, Lukacs K, Szwaja S, Bereczky, A. Alcohol–Diesel Fuel Combustion in the Compression Ignition Engine. Fuel 2015;154

[67] Sanli H, Canakci M, Alptekin E, Turkcan A, Ozsezen AN. Effects of Waste Frying Oil Based Methyl and Ethyl Ester Biodiesel Fuels on the Performance, Combustion and Emission Characteristics of a DI Diesel Engine.

[68] Shirneshan A. HC, CO, CO2 and NOx Emission Evaluation of a Diesel Engine Fueled with Waste Frying Oil Methyl Ester. Procedia-Social and Behavioral Sciences 2013;75 292-297.

intechopen.92569

Fuel 2015;159 179-187.

196-206.

Oil in Diesel Engines. Waste

[62] Mendez S, Kashdan JT, Bruneaux G, Thirouard B, Vangraefschepe F. Formation of Unburned Hydrocarbons in Low Temperature Diesel Combustion. SAE International Journal of Engines

Hill Education India.

2010;2(2) 205-225.

505-510.

**102**

## Mitigation of Emissions through Injection Strategies for C I Engine

*Jayashri N. Nair*

#### **Abstract**

Fuel conversion efficiency is high with diesel engines compared to petrol engines. However high emissions from diesel is a matter of concern and its mitigation paves way for scope of research. Exhaust gas recirculation is one of the method widely accepted to curb NOx emissions. Recently, split or multiple-injection strategy has been explored by researchers to precisely control the fuel injected per cycle and also to mitigate emissions. Present work reflects technical review of effect of injection strategies on performance, emissions and combustion on C.I. engine with diesel and biodiesel as fuel. Injection strategies like duration of injection, number of injections, the dwell period between two injections, quantity of injection, and multiple injections are analyzed for their influence on engine output and brake specific fuel consumption. Also their effect on emissions especially soot and NOx emission are reviewed. First the effect of injection strategies with diesel fuel is discussed followed by biodiesel.

**Keywords:** split injection, pilot injection, NOx emissions, Soot emissions, multiple injections

#### **1. Introduction**

The emissions from diesel engine are hazardous for human health as it causes respiratory problem in human beings. NOx emissions in combination with water and oxygen forms acid rain and is also responsible for the global warming. To improve the air quality in cities, regulations and control measures are taken to lower exhaust emissions from heavy vehicles. However, there are no much improvements mostly because of the difficulties in eliminating NOx and PM emissions. The need of an hour is simultaneous reduction of NOx and PM emissions from diesel engines. However it is challenging to control NOx and PM emission simultaneously. Optimizing injection timing through split injection can be a prominent solution to curb these emissions from engine. In split injection, the injection is divided into two or more pulses thus reducing the delay time. Major fraction of the combustion takes place during expansion stroke. NOx released from the engine is reduced as NOx is mostly formed in the premixed combustion. Investigations are done by many researchers to optimize the injection scheme and analyze the effect of split injections on combustion process and emissions.

Fuel injection timings influence the performance, emissions and the combustion part of the engine to a great extent. With the variation in the injection timings the state of air changes which effects the ignition delay. Occurrence of injection when the piston is far away from TDC reduces the temperature and pressure of the air

increasing the ignition delay. If injection starts when piston is nearer to TDC then temperature and pressure of the air will be higher which will decrease the ignition delay.

#### **2. Multiple injections**

Multiple fuel injections are employed, instead of single injection as in the conventional fuel injection system. The main injection provides maximum quantity of fuel per engine cycle. Pre injections can be one or more injections before the main injection. In many cases pre injections are termed as pilot injection. Pre injections can be employed when the compression stroke starts or can also be employed just before the main injection.

Multiple injections strategies involve following phenomena

1.Introducing fuel and dividing heat release rate (HRR).

2.Spatial distribution of fuel in combustion chamber.

3.Cooling effect of vaporizing fuel.

Linearity of the heat release rate is desirable for minimizing combustion noise. Pre injections helps in dividing the heat release process which eventually controls the combustion noise. Triple and quadruple injections also assists in minimizing combustion noise. Furthermore spatial fuel distribution is improved due to multiple injection strategies for effective usage of air for combustion in combustion chamber. Generally, this effect can lead to a reduction in particulate emissions at intermediate engine loads. The cooling effect associated with fuel vaporization lowers locally and globally the temperature of the gases contained in the combustion chamber. These phenomena can be applied to modify the rate of heat release in the early period of combustion by increasing the ignition there by allowing for a longer duration of mixing period and thus a more homogeneous fuel/air mixture. Correct use of cooling effect provides a valuable degree of freedom in the optimization of the noise, emission and fuel economy trade off. Multiple injections can be used for lower compression ratio engines with high EGR rates to mitigate the noise, emission and fuel economy trade off.

Post injections are injections after the main injection. It can occur immediately after the main injection or long time after the main injection and are also referred as after injections or late injections. Post injections are used to mitigate unburned hydrocarbons, particulate matter and for managing exhaust after treatment. Post injections can also be used with various EGR rates with single or double post injections to improve trade off relationship between NOx and HC, CO emissions.

#### **3. Effect of multiple injections on CI engine with diesel fuel**

#### **3.1 Effect of varying injection timing and multiple injections with diesel fuel**

In the fullness of time, clean, cheap and high torque diesel engines will be in demand. Controlled combustion is an essential part for an economic and clean diesel engine. Homogeneous air- fuel mixture plays a key role in effective combustion. Homogeneity of air-fuel mixture is affected by spatial fuel distribution in combustion chamber. Higher injection pressure, longer dwell time and increased

**105**

50% heat release.

for split injections.

*Mitigation of Emissions through Injection Strategies for C I Engine*

pilot fuel quantities can contribute to better mixing quality resulting in increased homogeneity of air-fuel mixture and optimum engine performance with low fuel consumption and soot emissions. However trade off does exists between NOx and soot emissions with more quantity of fuel in pilot injection. With changing quantity of fuel in first injection, reduction in both particulate emissions and BSFC was observed with single injection at fast rate [1]. Also optimizing the rate of injection for the lowest BSFC did not affect the pressure rise with single injection. However, it was evident that the quantity of fuel in the first injection effected the rate of incylinder pressure rise and NOx emissions also. Peak pressure reduced by more than 45% with split injection. The NOx emissions increased with more quantity of fuel in the first injection, whereas the particulate emissions decreased. High homogeneity of air fuel mixture leads to early burning of fuel in premixed combustion zone, increasing the in- cylinder combustion temperature resulting in high NOx emissions. Furthermore, split injection allows combustion to continue in the power stroke without any increase in soot emissions and thus utilizing the maximum charge. This indicates that pulsed injection may provide a means to reduce particu-

late emissions, and allow for reduced NOx from controlled pressure rise.

In another investigation, the injection pressure and the total quantity of fuel injected was kept constant but the ratio of amounts of fuel injected between two injection pulses were varied and also the time intervals between the two injections were varied [2]. With longer time interval between two injections and less quantity of fuel in the first injection, the in-line cylinder pressure decreased. Mixing of in-cylinder gas was enhanced by the second injection and the injection interval, which eliminated local high temperature areas resulting in reduced NOx emissions. Shorter time interval between two injections resulted in increased NOx emissions. The smoke emissions were high with less fuel quantity in the first injection due to the fact that large amount of oxygen was consumed during the first injection. More fuel was injected in the second injection which led to decreased combustion duration and resulted in increased smoke. However the smoke emission reduced with increased fuel quantity in the first injection. BSFC improved with less fuel quantity in second injection and shorter injection interval resulting in delayed crank angle at

Slight modification in the engine can assist in optimizing and treating the tradeoff between NOx and other emissions in combination with split injections. Design of double lobe cam with optimum split and dwell can influence reduction in engine emissions [3]. Use of double lobed cam showed reduction in NOx emission but showed reverse trend on CO and CO2 emissions. Split injection with 10° crank angle dwell using double lobed cam was optimum in effectively reducing the NOx emission by 14.1% and unburned hydro carbon by 11.8% without penalizing the soot and power. There was no considerable variation observed with SFC and brake thermal efficiency for the modified engine with split injection. In a similar study the amount of fuel injected was divided equally for pilot injection and main injection and was compared with the single injection [4]. Injection duration and injection rates were reduced. In case of split injection it was observed that the heat release rate reduced compared to the single injections resulting in lower NOx emissions. On retarding the split injection timing, soot emissions decreased compared to single injection. As the split injection timing was retarded, the larger size particles which forms the total particle volume and weight decreased, but HC and CO emissions were higher

Factors influencing emissions namely, ignition delay, adhered fuel, and squish are affected with the changes in the injection timings. Changes in injection timings affects the position of the piston, cylinder pressure and temperature at injection. Advancement of first stage injection timing to 80 °C BTDC, reduced BSFC to 20% [5].

*DOI: http://dx.doi.org/10.5772/intechopen.96483*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

Multiple injections strategies involve following phenomena

1.Introducing fuel and dividing heat release rate (HRR).

2.Spatial distribution of fuel in combustion chamber.

3.Cooling effect of vaporizing fuel.

tion delay.

**2. Multiple injections**

before the main injection.

and fuel economy trade off.

increasing the ignition delay. If injection starts when piston is nearer to TDC then temperature and pressure of the air will be higher which will decrease the igni-

Multiple fuel injections are employed, instead of single injection as in the conventional fuel injection system. The main injection provides maximum quantity of fuel per engine cycle. Pre injections can be one or more injections before the main injection. In many cases pre injections are termed as pilot injection. Pre injections can be employed when the compression stroke starts or can also be employed just

Linearity of the heat release rate is desirable for minimizing combustion noise. Pre injections helps in dividing the heat release process which eventually controls the combustion noise. Triple and quadruple injections also assists in minimizing combustion noise. Furthermore spatial fuel distribution is improved due to multiple injection strategies for effective usage of air for combustion in combustion chamber. Generally, this effect can lead to a reduction in particulate emissions at intermediate engine loads. The cooling effect associated with fuel vaporization lowers locally and globally the temperature of the gases contained in the combustion chamber. These phenomena can be applied to modify the rate of heat release in the early period of combustion by increasing the ignition there by allowing for a longer duration of mixing period and thus a more homogeneous fuel/air mixture. Correct use of cooling effect provides a valuable degree of freedom in the optimization of the noise, emission and fuel economy trade off. Multiple injections can be used for lower compression ratio engines with high EGR rates to mitigate the noise, emission

Post injections are injections after the main injection. It can occur immediately after the main injection or long time after the main injection and are also referred as after injections or late injections. Post injections are used to mitigate unburned hydrocarbons, particulate matter and for managing exhaust after treatment. Post injections can also be used with various EGR rates with single or double post injections to improve trade off relationship between NOx and HC, CO emissions.

**3. Effect of multiple injections on CI engine with diesel fuel**

**3.1 Effect of varying injection timing and multiple injections with diesel fuel**

In the fullness of time, clean, cheap and high torque diesel engines will be in demand. Controlled combustion is an essential part for an economic and clean diesel engine. Homogeneous air- fuel mixture plays a key role in effective combustion. Homogeneity of air-fuel mixture is affected by spatial fuel distribution in combustion chamber. Higher injection pressure, longer dwell time and increased

**104**

pilot fuel quantities can contribute to better mixing quality resulting in increased homogeneity of air-fuel mixture and optimum engine performance with low fuel consumption and soot emissions. However trade off does exists between NOx and soot emissions with more quantity of fuel in pilot injection. With changing quantity of fuel in first injection, reduction in both particulate emissions and BSFC was observed with single injection at fast rate [1]. Also optimizing the rate of injection for the lowest BSFC did not affect the pressure rise with single injection. However, it was evident that the quantity of fuel in the first injection effected the rate of incylinder pressure rise and NOx emissions also. Peak pressure reduced by more than 45% with split injection. The NOx emissions increased with more quantity of fuel in the first injection, whereas the particulate emissions decreased. High homogeneity of air fuel mixture leads to early burning of fuel in premixed combustion zone, increasing the in- cylinder combustion temperature resulting in high NOx emissions. Furthermore, split injection allows combustion to continue in the power stroke without any increase in soot emissions and thus utilizing the maximum charge. This indicates that pulsed injection may provide a means to reduce particulate emissions, and allow for reduced NOx from controlled pressure rise.

In another investigation, the injection pressure and the total quantity of fuel injected was kept constant but the ratio of amounts of fuel injected between two injection pulses were varied and also the time intervals between the two injections were varied [2]. With longer time interval between two injections and less quantity of fuel in the first injection, the in-line cylinder pressure decreased. Mixing of in-cylinder gas was enhanced by the second injection and the injection interval, which eliminated local high temperature areas resulting in reduced NOx emissions. Shorter time interval between two injections resulted in increased NOx emissions. The smoke emissions were high with less fuel quantity in the first injection due to the fact that large amount of oxygen was consumed during the first injection. More fuel was injected in the second injection which led to decreased combustion duration and resulted in increased smoke. However the smoke emission reduced with increased fuel quantity in the first injection. BSFC improved with less fuel quantity in second injection and shorter injection interval resulting in delayed crank angle at 50% heat release.

Slight modification in the engine can assist in optimizing and treating the tradeoff between NOx and other emissions in combination with split injections. Design of double lobe cam with optimum split and dwell can influence reduction in engine emissions [3]. Use of double lobed cam showed reduction in NOx emission but showed reverse trend on CO and CO2 emissions. Split injection with 10° crank angle dwell using double lobed cam was optimum in effectively reducing the NOx emission by 14.1% and unburned hydro carbon by 11.8% without penalizing the soot and power. There was no considerable variation observed with SFC and brake thermal efficiency for the modified engine with split injection. In a similar study the amount of fuel injected was divided equally for pilot injection and main injection and was compared with the single injection [4]. Injection duration and injection rates were reduced. In case of split injection it was observed that the heat release rate reduced compared to the single injections resulting in lower NOx emissions. On retarding the split injection timing, soot emissions decreased compared to single injection. As the split injection timing was retarded, the larger size particles which forms the total particle volume and weight decreased, but HC and CO emissions were higher for split injections.

Factors influencing emissions namely, ignition delay, adhered fuel, and squish are affected with the changes in the injection timings. Changes in injection timings affects the position of the piston, cylinder pressure and temperature at injection. Advancement of first stage injection timing to 80 °C BTDC, reduced BSFC to 20% [5]. The NOx and smoke emissions were improved with less quantity of fuel in the first injection and advancement of the injection from 80 °C BTDC to 100 °C BTDC. The CO emissions decreased when the first injection was advanced to 100 °C BTDC and the second injection was retarded over TDC.

#### **3.2 Effect of varying fuel injection pressure with diesel fuel**

Increasing fuel injection pressure is one of the best control techniques for optimizing combustion in diesel engines. High injection pressure promotes fine atomization of fuel, and uniform mixing of fuel and air thus, decreasing the combustion duration. Split injections coupled with higher pressure injection strategy tend to improve performance and emissions of engine. However combination of this strategy results in high NOx emissions. Also increased spray tip penetration due to high injection pressure results in spray impingement on wall and piston, leading to high HC emissions. In an investigation, conventional diesel fuel (Navy NATO F76) was compared with the new Navy hydro processed renewable diesel (HRD) fuel from algal sources, as well as the high cetane reference fuel nC16 (n-hexadecane CN¼100) [6]. It was observed that increasing fuel injection pressure shortened the ignition delay for all fuels. The combustion duration was longer for higher cetane number fuels for the same fuel pressure. This may be attributed to less premixing of fuel before the start of combustion. The author tried to study the relation between physical and chemical delay times with HRD and nC16. As injection pressure increases, the importance of chemical delay increases especially with lower cetane number. Split injections combined with exhaust gas recirculation (EGR) and higher injection pressure has the potential for diminishing NOx emissions and elevating the engine performance. The test was carried out on a V6 common rail direct injection engine [7]. The engine was turbocharged with variable turbine geometry turbochargers. The experiments were carried out at two different speeds 1500 rpm (35.1, 70.2 and 140 Nm) and 2000 rpm (43.3, 86.6 and 120 Nm) with injection pressure of 300, 430, 500, 600, 700 bar. The increase in injection pressure from 300 bar to 700 bar resulted in improved engine performance and emissions for all engine conditions. Increase in peak cylinder was observed for both EGR ON and EGR off conditions. In all engine test conditions the BSFC decreases as the injection pressure increases due to an improved fuel mixture and rapid combustion rate. However, these values were slightly increased when the engine operates with cooled EGR ON (1500 rpm, 35.1 Nm EGR OFF) at 600–700 bar injection pressure. This is strongly believed to be due to decreased combustion temperature and reduced oxygen availability for EGR ON. This results in higher fuel consumption. The reduced oxygen, higher inert in-cylinder gas and low boost pressure contributed to the poor combustion with cooled EGR ON. The specific fuel consumptions also decreased at both speeds. Reduction in CO emissions were observed with EGR off by 80% at 1500 rpm and 60% at 2000 rpm. The THC emissions also showed reduction with EGR OFF by approximately 70% at 1500 rpm and approximately 90% at 2000 rpm. However the NOx emissions increased 4 times at 1500 rpm especially at 70.2 Nm and by approximately 3 times at 2000 rpm especially at 43.4 Nm. The rate that NOx emissions increased as injection pressure rise was lower for EGR ON than for EGR OFF. Several factors that contribute to the fast combustion process in diesel engines are; fuel droplet size, penetration length, turbulence intensity, fuel evaporation, rate of combustion and ignition delay. However, all of these can contribute to the higher NOx formation. As expected, the EGR ON produces improved NOx reduction even at higher injection pressure for all engine test conditions. The use of cooled EGR produces lower peak in-cylinder pressure compared to cooled EGR OFF. This is due to a retarded combustion, a result of the reduced oxygen density and high heat

**107**

injection scheme.

*Mitigation of Emissions through Injection Strategies for C I Engine*

capacity of the gas mixture. In addition, the reduction of fuel burnt in the premixed combustion phase can be considered as a strong factor in lowering in-cylinder peak

**4.1 Effect of varying injection timing and multiple injections with biodiesel fuel**

The formation of NOx with biodiesel fuel blends, depend upon bulk modulus of biodiesel, oxygen content of biodiesel, cetane number, saturated fatty acid content and engine operating conditions. Varying injection timing has proved to be a potential technique in curbing NOx emissions in case of biodiesel. During the premixed combustion phase, NOx is formed in lean flame region. Optimization of combustion timing and regulating the quantity of fuel burned in premixed phase

Coconut biodiesel with 20% and 50% proportions by volume in diesel were tested with single, double and triple injections [8]. In double injection the fuel mass was distributed equally for both injections while in triple injection 33.33% of mass was distributed in each injections with a dwell time of 1.3 ms in both double

Advanced start of injection timings increased brake thermal efficiency for all fuels. This may be due to longer ignition delay which promotes better mixing of fuel and air resulting in efficient combustion. Also in case of advanced injections the peak

efficiency decreased with the increase in the biodiesel blends. For all injection timings and schemes the BSFC was higher for biodiesel blends in comparison to diesel. Due to the lower heating value of the biodiesel more mass is needed to produce the same power output. Also it is noted that with increase in split injections the BSFC increases for all the fuels. This may be due to the longer duration of combustion process with the higher number of split injections which increases the heat loss and reduces the peak pressure which reduces the work output. NOx is increased with increase in the advancement of injections due to occurrence of peak pressure around TDC which results in higher temperature. Biodiesel blends showed less NOx emissions than diesel for all split injections and single injection. This may be due to higher cetane number and lower calorific values of the blends which helps in reducing the rate of heat release. With B50 blend simultaneous reduction of NOx and smoke emission is possible with retardation in injection timing with triple

In another investigation effect of injection timing on NOx emissions with ultralow sulfur diesel fuel and biodiesel blends (name not specified) were studied [9]. At low load conditions biodiesel blends were found to produce slightly lower NOx emissions than the neat diesel fuel. However at higher loads the NOx emissions were higher for both single and double injection conditions. Overall, biodiesel diesel blends and the baseline diesel fuel had very similar heat release rate profiles. When the injection timing was retarded it proved effective for single injection conditions than using pilot injection with retarded main injection in reducing the NOx emissions. Under the low load condition, the pilot injection strategy led to substantially reduced NOx emissions. Similar effect of retardation on NOx emission is reported

using Pongmia methyl ester [10]. The injection timing was retarded by 4o

all the blends with retarded injection timing.

blends. All the blends (B20, B40, B60, and B80) showed lower NOx emissions and HC emissions with retarded injection timing. The smoke emissions also reduced for

BTDC to 2o

BTDC brake thermal

ATDC.

for the

**4. Effect of multiple injections on CI engine with biodiesel fuel**

*DOI: http://dx.doi.org/10.5772/intechopen.96483*

can significantly reduce the NOx emissions.

and triple injection case. Injection timing was varied from 12o

pressure was attained at TDC. With single injection at 12o

pressure.

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**3.2 Effect of varying fuel injection pressure with diesel fuel**

the second injection was retarded over TDC.

The NOx and smoke emissions were improved with less quantity of fuel in the first injection and advancement of the injection from 80 °C BTDC to 100 °C BTDC. The CO emissions decreased when the first injection was advanced to 100 °C BTDC and

Increasing fuel injection pressure is one of the best control techniques for optimizing combustion in diesel engines. High injection pressure promotes fine atomization of fuel, and uniform mixing of fuel and air thus, decreasing the combustion duration. Split injections coupled with higher pressure injection strategy tend to improve performance and emissions of engine. However combination of this strategy results in high NOx emissions. Also increased spray tip penetration due to high injection pressure results in spray impingement on wall and piston, leading to high HC emissions. In an investigation, conventional diesel fuel (Navy NATO F76) was compared with the new Navy hydro processed renewable diesel (HRD) fuel from algal sources, as well as the high cetane reference fuel nC16 (n-hexadecane CN¼100) [6]. It was observed that increasing fuel injection pressure shortened the ignition delay for all fuels. The combustion duration was longer for higher cetane number fuels for the same fuel pressure. This may be attributed to less premixing of fuel before the start of combustion. The author tried to study the relation between physical and chemical delay times with HRD and nC16. As injection pressure increases, the importance of chemical delay increases especially with lower cetane number. Split injections combined with exhaust gas recirculation (EGR) and higher injection pressure has the potential for diminishing NOx emissions and elevating the engine performance. The test was carried out on a V6 common rail direct injection engine [7]. The engine was turbocharged with variable turbine geometry turbochargers. The experiments were carried out at two different speeds 1500 rpm (35.1, 70.2 and 140 Nm) and 2000 rpm (43.3, 86.6 and 120 Nm) with injection pressure of 300, 430, 500, 600, 700 bar. The increase in injection pressure from 300 bar to 700 bar resulted in improved engine performance and emissions for all engine conditions. Increase in peak cylinder was observed for both EGR ON and EGR off conditions. In all engine test conditions the BSFC decreases as the injection pressure increases due to an improved fuel mixture and rapid combustion rate. However, these values were slightly increased when the engine operates with cooled EGR ON (1500 rpm, 35.1 Nm EGR OFF) at 600–700 bar injection pressure. This is strongly believed to be due to decreased combustion temperature and reduced oxygen availability for EGR ON. This results in higher fuel consumption. The reduced oxygen, higher inert in-cylinder gas and low boost pressure contributed to the poor combustion with cooled EGR ON. The specific fuel consumptions also decreased at both speeds. Reduction in CO emissions were observed with EGR off by 80% at 1500 rpm and 60% at 2000 rpm. The THC emissions also showed reduction with EGR OFF by approximately 70% at 1500 rpm and approximately 90% at 2000 rpm. However the NOx emissions increased 4 times at 1500 rpm especially at 70.2 Nm and by approximately 3 times at 2000 rpm especially at 43.4 Nm. The rate that NOx emissions increased as injection pressure rise was lower for EGR ON than for EGR OFF. Several factors that contribute to the fast combustion process in diesel engines are; fuel droplet size, penetration length, turbulence intensity, fuel evaporation, rate of combustion and ignition delay. However, all of these can contribute to the higher NOx formation. As expected, the EGR ON produces improved NOx reduction even at higher injection pressure for all engine test conditions. The use of cooled EGR produces lower peak in-cylinder pressure compared to cooled EGR OFF. This is due to a retarded combustion, a result of the reduced oxygen density and high heat

**106**

capacity of the gas mixture. In addition, the reduction of fuel burnt in the premixed combustion phase can be considered as a strong factor in lowering in-cylinder peak pressure.

### **4. Effect of multiple injections on CI engine with biodiesel fuel**

#### **4.1 Effect of varying injection timing and multiple injections with biodiesel fuel**

The formation of NOx with biodiesel fuel blends, depend upon bulk modulus of biodiesel, oxygen content of biodiesel, cetane number, saturated fatty acid content and engine operating conditions. Varying injection timing has proved to be a potential technique in curbing NOx emissions in case of biodiesel. During the premixed combustion phase, NOx is formed in lean flame region. Optimization of combustion timing and regulating the quantity of fuel burned in premixed phase can significantly reduce the NOx emissions.

Coconut biodiesel with 20% and 50% proportions by volume in diesel were tested with single, double and triple injections [8]. In double injection the fuel mass was distributed equally for both injections while in triple injection 33.33% of mass was distributed in each injections with a dwell time of 1.3 ms in both double and triple injection case. Injection timing was varied from 12o BTDC to 2o ATDC. Advanced start of injection timings increased brake thermal efficiency for all fuels. This may be due to longer ignition delay which promotes better mixing of fuel and air resulting in efficient combustion. Also in case of advanced injections the peak pressure was attained at TDC. With single injection at 12o BTDC brake thermal efficiency decreased with the increase in the biodiesel blends. For all injection timings and schemes the BSFC was higher for biodiesel blends in comparison to diesel. Due to the lower heating value of the biodiesel more mass is needed to produce the same power output. Also it is noted that with increase in split injections the BSFC increases for all the fuels. This may be due to the longer duration of combustion process with the higher number of split injections which increases the heat loss and reduces the peak pressure which reduces the work output. NOx is increased with increase in the advancement of injections due to occurrence of peak pressure around TDC which results in higher temperature. Biodiesel blends showed less NOx emissions than diesel for all split injections and single injection. This may be due to higher cetane number and lower calorific values of the blends which helps in reducing the rate of heat release. With B50 blend simultaneous reduction of NOx and smoke emission is possible with retardation in injection timing with triple injection scheme.

In another investigation effect of injection timing on NOx emissions with ultralow sulfur diesel fuel and biodiesel blends (name not specified) were studied [9]. At low load conditions biodiesel blends were found to produce slightly lower NOx emissions than the neat diesel fuel. However at higher loads the NOx emissions were higher for both single and double injection conditions. Overall, biodiesel diesel blends and the baseline diesel fuel had very similar heat release rate profiles. When the injection timing was retarded it proved effective for single injection conditions than using pilot injection with retarded main injection in reducing the NOx emissions. Under the low load condition, the pilot injection strategy led to substantially reduced NOx emissions. Similar effect of retardation on NOx emission is reported using Pongmia methyl ester [10]. The injection timing was retarded by 4o for the blends. All the blends (B20, B40, B60, and B80) showed lower NOx emissions and HC emissions with retarded injection timing. The smoke emissions also reduced for all the blends with retarded injection timing.

#### **4.2 Effect of varying quantity of fuel in the injections with biodiesel fuel**

Varying quantity of fuel in multiple injections affects the performance and emission of the engine [11]. For single injection the quantity of fuel discharged was 10 mg whereas for multiple injections the fuel quantity was divided in equal proportion for both injections in first test (5 mg + 5 mg) and 3 mg in first injection and 7 mg in second injection for second test. With single injection start of energizing was arranged at TDC, 10o BTDC, 20o BTDC whereas for multiple injections it was arranged at TDC, 10o BTDC, 20o BTDC, 30 o BTDC. In case of single injection test, injection pressures varied to 60 MPa and 120 MPa. For single injection as the advancement increased the peak combustion pressure and peak heat release rate increased. However at 30 o BTDC these values dropped. This can be attributed to the inline cylinder pressure and temperature which were lower during the spray and resulted in long ignition delay. Also the fuel injected dispersed in the squish and crevice regions. Soot emissions were lower with higher injection pressure. Peak NOx emission was reported at 20o BTDC for 60 MPa and 15o BTDC for 120 MPa. For injection timings below these values, the NOx emissions continuously decreased. The increase in NOx emission was attributed to the proper mixing of the spray with air. The decrease in the NOx emission at less injection timings was attributed to the incomplete combustion. CO and HC emissions increased with the advancement of the injection timings. In case of multiple injection at 120 MPa with 20o BTDC pilot injection and main injection at TDC showed lower heat release rates than single injection. Due to shorter ignition delay the second combustion resumed quickly than the first combustion consequently the temperature of combustion chamber increased for first combustion. During the first injection the consumption of oxygen will be maximum and so the rate of heat release was higher for the first injection as compared to second injection. It was observed that with shorter ignition timings supported less soot, HC and CO emissions but NOx increased.

#### **4.3 Effect of varying fuel injection pressure with biodiesel fuel**

Proportion of biodiesel in diesel influences the injection spray pattern [12]. Due to the increased fuel line pressure, with increase of biodiesel in the blends the penetration distance of the spray also increases. Simulation results reports that the probability of wall impingement is more with higher blends. The ignition delay period decreased with all biodiesel blends resulting in less rate of pressure rise. This may attributed to higher cetane number of the biodiesel. Reduction in torque is noted with B100 at rated load. However for rest of the blends there was no significant change in the torque. HC, CO and smoke emissions decreased with all blends while NOx emissions increased in the range of 1.4–22.8% with all biodiesel. This may be attributed to the oxygenated fuel and automatic advance in dynamic injection timing. It was concluded that B15 was the optimum blend with respect to no wall impingement and NOx emissions.

Tests were conducted to evaluate best injection timing and injection pressure with Honge methyl ester and cotton seed oil methyl ester [13]. A toroidal re-entrant combustion chamber of the engine was selected. The injection pressure was varied from 220 to 260 bar and the injection timing of 19°, 23° and 27° BTDC was executed at compression ratio of 17.5. At injection timing of 19° BTDC and injection pressure of 240 bar biodiesel showed best BTE. Honge methyl ester showed better performance compared to cotton seed methyl ester. Further the injector of six holes with 0.2 mm orifice diameter yielded better results compared to 0.25 mm and 0.3 mm orifice diameter. It showed lower brake thermal efficiency for the biodiesel fuel. Smaller orifices have shorter ignition delay which reduces heat loss and time loss

**109**

*Mitigation of Emissions through Injection Strategies for C I Engine*

resulting in higher brake thermal efficiency. The NOx emissions were lower whereas the HC and CO emissions were higher with 0.2 mm orifice diameter for the bio-

Tests were conducted for three injection pressures 180, 200, 220 bar for four different loads at 2200 rpm with methanol blended diesel from 0–15% with an increment of 5% [14]. The original pressure of the engine was 200 bar. It was reported that at decreased pressure of 180 bar the heat release rate, combustion efficiency, NOx emissions, CO2 emissions and peak cylinder pressure decreased whereas CO, HC and smoke number increased. On the other hand, smoke number, unburned hydrocarbon, and carbon monoxide emissions reduced with 220 bar, and peak cylinder pressure, heat release rate, combustion efficiency, nitrogen oxides and

Exhaust emissions and BSFC with biodiesel blends (Name not specified) on diesel engine with different injection pressures at four different loads were studied [15]. All the biodiesel blends showed less CO emissions at lighter loads and high emissions at full loads. It was also noted that the CO emissions and HC emissions decreased with the increase in the biodiesel percentage. At 50 kPa constant load the CO emissions, HC emissions and smoke opacity decreased with the increasing injection pressure. As the engine load increased the NOx emissions increased for all the blends. There was a rise in the NOx emissions with the increase in the percentage of the biodiesel. With the increase in injection pressures, blends with high percent-

Optimization of injection pressure and compression ratio can improve engine performance [16]. The pressures selected were (150 bar, 200 bar, 250 bar) and compression ratios selected for the test were 16, 17, and 18. On comparing all the combination of pressures and compression ratio it was found that at 250 bar injection pressure and 18 compression ratio gave the highest brake thermal efficiency and lowest brake specific fuel consumption for the Jatropha methyl ester. However the HC and exhaust temperature increased with this combination whereas the smoke and CO emissions reduced. No change was observed with the NOx emissions at higher pressures. For all the combination of pressure and compression ratio the Jatropha methyl ester showed better results than neat diesel both in case of perfor-

In a similar study higher BSFC with linolenic linseed oil methyl ester at higher pressure of 240 bar was reported [17]. Moreover the thermal efficiency improved at 240 bar with some increased emission in NOx. This may due to the changes in the shape of fuel spray which results in shortening of the length of the spray resulting in higher dispersion and higher spray tip penetration. This ultimately lead to better combustion.The ignition delay was reported to be lower at higher injection pressures as compared to diesel and at full load the peak pressure was the highest. The combustion analysis shows that, the ignition delay is lower at higher injection pressures compared to diesel. Peak pressures increased with the increase in the load for all injection pressures for the biodiesel as well as diesel. Smoke was lower than

Fuel reactivity influenced simultaneous reduction in NOx and soot emissions and increased dwell period, reduced particulate matter for diesel fuel. Multiple injection strategies are effective only with certain adjustment with injection rate as well as the duration of dwell period. Quantity of fuel in the first and second injection also greatly influences brake thermal efficiency and smoke emissions.

*DOI: http://dx.doi.org/10.5772/intechopen.96483*

carbon dioxide emissions increased at all loads.

age of biodiesel showed less BSFC.

mance as well as emissions.

diesel at all loads for the biodiesel.

**5. Conclusion**

diesel fuel.

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

ing was arranged at TDC, 10o

was arranged at TDC, 10o

increased. However at 30 o

NOx emission was reported at 20o

no wall impingement and NOx emissions.

**4.2 Effect of varying quantity of fuel in the injections with biodiesel fuel**

BTDC, 20o

BTDC, 20o

Varying quantity of fuel in multiple injections affects the performance and emission of the engine [11]. For single injection the quantity of fuel discharged was 10 mg whereas for multiple injections the fuel quantity was divided in equal proportion for both injections in first test (5 mg + 5 mg) and 3 mg in first injection and 7 mg in second injection for second test. With single injection start of energiz-

BTDC, 30 o

test, injection pressures varied to 60 MPa and 120 MPa. For single injection as the advancement increased the peak combustion pressure and peak heat release rate

the inline cylinder pressure and temperature which were lower during the spray and resulted in long ignition delay. Also the fuel injected dispersed in the squish and crevice regions. Soot emissions were lower with higher injection pressure. Peak

injection timings below these values, the NOx emissions continuously decreased. The increase in NOx emission was attributed to the proper mixing of the spray with air. The decrease in the NOx emission at less injection timings was attributed to the incomplete combustion. CO and HC emissions increased with the advancement of the injection timings. In case of multiple injection at 120 MPa with 20o

pilot injection and main injection at TDC showed lower heat release rates than single injection. Due to shorter ignition delay the second combustion resumed quickly than the first combustion consequently the temperature of combustion chamber increased for first combustion. During the first injection the consumption of oxygen will be maximum and so the rate of heat release was higher for the first injection as compared to second injection. It was observed that with shorter ignition

Proportion of biodiesel in diesel influences the injection spray pattern [12]. Due to the increased fuel line pressure, with increase of biodiesel in the blends the penetration distance of the spray also increases. Simulation results reports that the probability of wall impingement is more with higher blends. The ignition delay period decreased with all biodiesel blends resulting in less rate of pressure rise. This may attributed to higher cetane number of the biodiesel. Reduction in torque is noted with B100 at rated load. However for rest of the blends there was no significant change in the torque. HC, CO and smoke emissions decreased with all blends while NOx emissions increased in the range of 1.4–22.8% with all biodiesel. This may be attributed to the oxygenated fuel and automatic advance in dynamic injection timing. It was concluded that B15 was the optimum blend with respect to

Tests were conducted to evaluate best injection timing and injection pressure with Honge methyl ester and cotton seed oil methyl ester [13]. A toroidal re-entrant combustion chamber of the engine was selected. The injection pressure was varied from 220 to 260 bar and the injection timing of 19°, 23° and 27° BTDC was executed at compression ratio of 17.5. At injection timing of 19° BTDC and injection pressure of 240 bar biodiesel showed best BTE. Honge methyl ester showed better performance compared to cotton seed methyl ester. Further the injector of six holes with 0.2 mm orifice diameter yielded better results compared to 0.25 mm and 0.3 mm orifice diameter. It showed lower brake thermal efficiency for the biodiesel fuel. Smaller orifices have shorter ignition delay which reduces heat loss and time loss

timings supported less soot, HC and CO emissions but NOx increased.

**4.3 Effect of varying fuel injection pressure with biodiesel fuel**

BTDC for 60 MPa and 15o

BTDC whereas for multiple injections it

BTDC these values dropped. This can be attributed to

BTDC. In case of single injection

BTDC for 120 MPa. For

BTDC

**108**

resulting in higher brake thermal efficiency. The NOx emissions were lower whereas the HC and CO emissions were higher with 0.2 mm orifice diameter for the biodiesel fuel.

Tests were conducted for three injection pressures 180, 200, 220 bar for four different loads at 2200 rpm with methanol blended diesel from 0–15% with an increment of 5% [14]. The original pressure of the engine was 200 bar. It was reported that at decreased pressure of 180 bar the heat release rate, combustion efficiency, NOx emissions, CO2 emissions and peak cylinder pressure decreased whereas CO, HC and smoke number increased. On the other hand, smoke number, unburned hydrocarbon, and carbon monoxide emissions reduced with 220 bar, and peak cylinder pressure, heat release rate, combustion efficiency, nitrogen oxides and carbon dioxide emissions increased at all loads.

Exhaust emissions and BSFC with biodiesel blends (Name not specified) on diesel engine with different injection pressures at four different loads were studied [15]. All the biodiesel blends showed less CO emissions at lighter loads and high emissions at full loads. It was also noted that the CO emissions and HC emissions decreased with the increase in the biodiesel percentage. At 50 kPa constant load the CO emissions, HC emissions and smoke opacity decreased with the increasing injection pressure. As the engine load increased the NOx emissions increased for all the blends. There was a rise in the NOx emissions with the increase in the percentage of the biodiesel. With the increase in injection pressures, blends with high percentage of biodiesel showed less BSFC.

Optimization of injection pressure and compression ratio can improve engine performance [16]. The pressures selected were (150 bar, 200 bar, 250 bar) and compression ratios selected for the test were 16, 17, and 18. On comparing all the combination of pressures and compression ratio it was found that at 250 bar injection pressure and 18 compression ratio gave the highest brake thermal efficiency and lowest brake specific fuel consumption for the Jatropha methyl ester. However the HC and exhaust temperature increased with this combination whereas the smoke and CO emissions reduced. No change was observed with the NOx emissions at higher pressures. For all the combination of pressure and compression ratio the Jatropha methyl ester showed better results than neat diesel both in case of performance as well as emissions.

In a similar study higher BSFC with linolenic linseed oil methyl ester at higher pressure of 240 bar was reported [17]. Moreover the thermal efficiency improved at 240 bar with some increased emission in NOx. This may due to the changes in the shape of fuel spray which results in shortening of the length of the spray resulting in higher dispersion and higher spray tip penetration. This ultimately lead to better combustion.The ignition delay was reported to be lower at higher injection pressures as compared to diesel and at full load the peak pressure was the highest. The combustion analysis shows that, the ignition delay is lower at higher injection pressures compared to diesel. Peak pressures increased with the increase in the load for all injection pressures for the biodiesel as well as diesel. Smoke was lower than diesel at all loads for the biodiesel.

#### **5. Conclusion**

Fuel reactivity influenced simultaneous reduction in NOx and soot emissions and increased dwell period, reduced particulate matter for diesel fuel. Multiple injection strategies are effective only with certain adjustment with injection rate as well as the duration of dwell period. Quantity of fuel in the first and second injection also greatly influences brake thermal efficiency and smoke emissions.

Split injections with biodiesel has mitigated NOx emissions but at the cost of engine performances. Split injections with higher injection pressure reduced soot, HC and CO emissions at the cost of NOx emissions. Higher injection pressures with biodiesel did not show consistent results with NOx emissions however when coupled with EGR strategies, it has depicted significant reduction in NOx emissions.

### **Author details**

Jayashri N. Nair VNR Vignana Jyothi Institute of Engineering and Technology, Hyderabad, India

\*Address all correspondence to: jayashri@vnrvjiet.in

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

**111**

*Mitigation of Emissions through Injection Strategies for C I Engine*

[8] How, H.G., Masjuki, H.H., Kalam, M.A. and Teoh, Y.H., Influence of Injection Timing and Split Injection Strategies on Performance, Emissions, and Combustion Characteristics of Diesel Engine Fuelled with Biodiesel Blended Fuels. Fuel. 2018; 213: 106-114.

DOI: 10.1016/j.fuel.2017.10.102

[9] Zhang, Y. and Boehman, A.L., Impact of Biodiesel on NOx Emissions in a Common Rail Direct Injection Diesel Engine. Energy & Fuels. 2007: 2003-2012. DOI: 10.1021/ef0700073

[10] Suryawanshi, J.G. and Deshpande, N.V., Effect of Injection Timing Retard on Emissions and Performance of a Pongamia Oil Methyl Ester Fuelled CI Engine (No. 2005-01-3677). SAE Technical Paper 24 Oct 2005; DOI:

10.4271/2005-01-3677.

fuel.2010.08.029

[11] Park SH, Youn IM, Lee CS., Influence of Ethanol Blends on the Combustion Performance and Exhaust Emission Characteristics of a Fourcylinder Diesel Engine at various Engine Loads and Injection Timings. Fuel. 2011; 90:748-55. DOI: 10.1016/j.

[12] Lahane S, Subramanian KA., Effect of Different Percentages of Biodiesel–diesel Blends on Injection, Spray, Combustion, Performance, and Emission Characteristics of a Diesel Engine. Fuel.2015; 139: 537-45. DOI:

[13] Khandal, S.V., Banapurmath, N.R. and Gaitonde, V.N., Different Injection Strategies to Enhance the Performance of Diesel Engine Powered with Biodiesel Fuels. European Journal of Sustainable Development Research. 2017; 1-5. DOI:

[14] Canakci, M., Sayin, C., Ozsezen, A.N. and Turkcan, A., Effect of Injection Pressure on the Combustion,

10.1016/j.fuel.2014.09.036

10.20897/ejosdr.201715

*DOI: http://dx.doi.org/10.5772/intechopen.96483*

[1] Nehmer DA, Reitz RD., Measurement of the Effect of Injection Rate and Split Injections on Diesel Engine Soot and NOx emissions. SAE Transactions,

[2] Akiba S, Keiya N, Xinyun Z., Effect of Split Injection Pattern on Combustion and Emission Characteristics of DI Diesel Engine. Marine Engineering. Journal of the Japan Institution of Marine Engineering. 2011; 46:441-47.

**References**

pp.1030-1041, (1994)

DOI: 10.5988/jime.46.441

Technology.2015; 10:552-70.

ef700537w

energy.2004.05.009

DOI: 10.1115/1.4027408

10.4271/2009-24-0049.

[3] Krishna P, Babu AK, Singh AP, Raj AA. Reduction of NOx in a Diesel Engine Using Split Injection Approach. Journal of Engineering Science and

[4] Kim, M.Y., Yoon, S.H. and Lee, C.S., Impact of Split Injection Strategy on the Exhaust Emissions and Soot Particulates from a Compression Ignition Engine Fuelled with Neat Biodiesel. Energy & Fuels. 2008; 22:1260-1265. DOI: 10.1021/

[5] Iwazaki, K., Amagai, K. and Arai, M., Improvement of Fuel Economy of an Indirect Injection (IDI) Diesel Engine with Two-stage Injection. Energy. 2005; 30: 447-459. DOI: 10.1016/j.

[6] Cowart, Jim, Dianne Luning Prak, and Len Hamilton. The Effects of Fuel Injection Pressure and Fuel Type on the Combustion Characteristics of a Diesel Engine. In ASME 2014 Internal Combustion Engine Division Fall Technical Conference, American Society of Mechanical Engineers. 2014.

[7] Abdullah N, Tsolakis A, Rounce P, Wyszinsky M, Xu H, Mamat R. Effect of injection pressure with split Injection in a V6 diesel engine. SAE Paper NO. 2009-24-0049. DOI:

*Mitigation of Emissions through Injection Strategies for C I Engine DOI: http://dx.doi.org/10.5772/intechopen.96483*

#### **References**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

Split injections with biodiesel has mitigated NOx emissions but at the cost of engine performances. Split injections with higher injection pressure reduced soot, HC and CO emissions at the cost of NOx emissions. Higher injection pressures with biodiesel did not show consistent results with NOx emissions however when coupled with EGR strategies, it has depicted significant reduction in NOx emissions.

**110**

**Author details**

Jayashri N. Nair

VNR Vignana Jyothi Institute of Engineering and Technology, Hyderabad, India

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium,

\*Address all correspondence to: jayashri@vnrvjiet.in

provided the original work is properly cited.

[1] Nehmer DA, Reitz RD., Measurement of the Effect of Injection Rate and Split Injections on Diesel Engine Soot and NOx emissions. SAE Transactions, pp.1030-1041, (1994)

[2] Akiba S, Keiya N, Xinyun Z., Effect of Split Injection Pattern on Combustion and Emission Characteristics of DI Diesel Engine. Marine Engineering. Journal of the Japan Institution of Marine Engineering. 2011; 46:441-47. DOI: 10.5988/jime.46.441

[3] Krishna P, Babu AK, Singh AP, Raj AA. Reduction of NOx in a Diesel Engine Using Split Injection Approach. Journal of Engineering Science and Technology.2015; 10:552-70.

[4] Kim, M.Y., Yoon, S.H. and Lee, C.S., Impact of Split Injection Strategy on the Exhaust Emissions and Soot Particulates from a Compression Ignition Engine Fuelled with Neat Biodiesel. Energy & Fuels. 2008; 22:1260-1265. DOI: 10.1021/ ef700537w

[5] Iwazaki, K., Amagai, K. and Arai, M., Improvement of Fuel Economy of an Indirect Injection (IDI) Diesel Engine with Two-stage Injection. Energy. 2005; 30: 447-459. DOI: 10.1016/j. energy.2004.05.009

[6] Cowart, Jim, Dianne Luning Prak, and Len Hamilton. The Effects of Fuel Injection Pressure and Fuel Type on the Combustion Characteristics of a Diesel Engine. In ASME 2014 Internal Combustion Engine Division Fall Technical Conference, American Society of Mechanical Engineers. 2014. DOI: 10.1115/1.4027408

[7] Abdullah N, Tsolakis A, Rounce P, Wyszinsky M, Xu H, Mamat R. Effect of injection pressure with split Injection in a V6 diesel engine. SAE Paper NO. 2009-24-0049. DOI: 10.4271/2009-24-0049.

[8] How, H.G., Masjuki, H.H., Kalam, M.A. and Teoh, Y.H., Influence of Injection Timing and Split Injection Strategies on Performance, Emissions, and Combustion Characteristics of Diesel Engine Fuelled with Biodiesel Blended Fuels. Fuel. 2018; 213: 106-114. DOI: 10.1016/j.fuel.2017.10.102

[9] Zhang, Y. and Boehman, A.L., Impact of Biodiesel on NOx Emissions in a Common Rail Direct Injection Diesel Engine. Energy & Fuels. 2007: 2003-2012. DOI: 10.1021/ef0700073

[10] Suryawanshi, J.G. and Deshpande, N.V., Effect of Injection Timing Retard on Emissions and Performance of a Pongamia Oil Methyl Ester Fuelled CI Engine (No. 2005-01-3677). SAE Technical Paper 24 Oct 2005; DOI: 10.4271/2005-01-3677.

[11] Park SH, Youn IM, Lee CS., Influence of Ethanol Blends on the Combustion Performance and Exhaust Emission Characteristics of a Fourcylinder Diesel Engine at various Engine Loads and Injection Timings. Fuel. 2011; 90:748-55. DOI: 10.1016/j. fuel.2010.08.029

[12] Lahane S, Subramanian KA., Effect of Different Percentages of Biodiesel–diesel Blends on Injection, Spray, Combustion, Performance, and Emission Characteristics of a Diesel Engine. Fuel.2015; 139: 537-45. DOI: 10.1016/j.fuel.2014.09.036

[13] Khandal, S.V., Banapurmath, N.R. and Gaitonde, V.N., Different Injection Strategies to Enhance the Performance of Diesel Engine Powered with Biodiesel Fuels. European Journal of Sustainable Development Research. 2017; 1-5. DOI: 10.20897/ejosdr.201715

[14] Canakci, M., Sayin, C., Ozsezen, A.N. and Turkcan, A., Effect of Injection Pressure on the Combustion,

**Chapter 7**

**Abstract**

**1. Introduction**

**113**

Combustion and Emissions of

Gasoline Compression Ignition

A gasoline compression ignition (GCI) engine was proposed to be the next generation internal combustion engine for gasoline. The effect of exhaust gas recirculation (EGR) and intake boosting on combustion and emissions of GCI engine fueled with gasoline-biodiesel blends by partially premixed compression ignition (PPCI) combustions are investigated in this study. Tests were conducted on a single-cylinder direct-injection CI engine, with 5% by volume proportion of biodiesel in gasoline fuel blends. Engine control parameters (EGR rate, intake boosting rate, and various injection strategies) were adjusted to investigate their influences on combustion and emissions of this GCI engine. It is found that changes in EGR rate, intake boosting pressure and injection strategies affect on ignition delay, maximum pressure rise rate and thermal efficiency which is closely tied to HC, CO,

The stricter limitations of vehicle emission regulation especially for compression

ignition (CI) engines with petroleum diesel fuel motivates many researchers to explore the utilization of low volatile and alternative fuels for CI engines to obtain high efficiency, but produce lower emissions or so-called low temperature combustion (LTC). Due to the low volatility and short ignition delay of diesel fuel, CI engines produce high nitrogen oxides (NOx) and soot emissions. To obtain LTC combustion, which is increase engine efficiency, and improve exhaust emissions, a variety of combustion methods in CI engines have been investigated such as homogeneous charge compression ignition (HCCI), PPCI, multiple premixed compression ignition (MPCI), etc. HCCI engine is one method that potentially to achieve an advanced LTC, which possible to produce low particulate matter (PM) and NOx emissions to replace conventional diesel engine combustion [1]. However, many technical problems of HCCI engine strategy must be solved before released to the market. The maximum load limitation due to the surplus of pressure rise rate (PRR) and engine knocking phenomenon [2, 3] have been major obstacle as long as HCCI

Engine Fuelled with

*Yanuandri Putrasari and Ocktaeck Lim*

NOx and smoke emissions, respectively.

**Keywords:** gasoline, fuel, GCI, thermal efficiency, emission

combustion influenced by fuel type and air fuel mixture quality.

Gasoline-Biodiesel Blends

Performance, and Emission Characteristics of a Diesel Engine Fuelled with Methanol-blended Diesel Fuel. Energy & Fuels. 2009; 23:2908- 2920. DOI: doi.org/10.1021/ef900060s

[15] Gumus, M., Sayin, C., & Canakci, M. The Impact of Fuel Injection Pressure on the Exhaust Emissions of a Direct Injection Diesel Engine Fuelled with Biodiesel–diesel Fuel Blends. Fuel. 2012; 95: 486-494. DOI: 10.1016/j. fuel.2011.11.020

[16] Jindal, S., Nandwana, B.P., Rathore, N.S. and Vashistha, V., Experimental Investigation of the Effect of Compression Ratio and Injection Pressure in a Direct Injection Diesel Engine Running on Jatropha Methyl Ester. Applied Thermal Engineering. 2010; 30:442-448. DOI: 10.1016/j. applthermaleng.2009.10.004

[17] Puhan, S., Jegan, R., Balasubbramanian, K. and Nagarajan, G., Effect of Injection Pressure on Performance, Emission and Combustion Characteristics of High Linolenic Linseed Oil Methyl Ester in a DI Diesel Engine. Renewable energy. 2009; 34:1227-1233. DOI: 10.1016/j. renene.2008.10.001

#### **Chapter 7**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

Performance, and Emission Characteristics of a Diesel Engine Fuelled with Methanol-blended Diesel Fuel. Energy & Fuels. 2009; 23:2908- 2920. DOI: doi.org/10.1021/ef900060s

fuel.2011.11.020

[15] Gumus, M., Sayin, C., & Canakci, M. The Impact of Fuel Injection Pressure on the Exhaust Emissions of a Direct Injection Diesel Engine Fuelled with Biodiesel–diesel Fuel Blends. Fuel. 2012; 95: 486-494. DOI: 10.1016/j.

[16] Jindal, S., Nandwana, B.P., Rathore, N.S. and Vashistha, V.,

[17] Puhan, S., Jegan, R.,

renene.2008.10.001

Experimental Investigation of the Effect of Compression Ratio and Injection Pressure in a Direct Injection Diesel Engine Running on Jatropha Methyl Ester. Applied Thermal Engineering. 2010; 30:442-448. DOI: 10.1016/j. applthermaleng.2009.10.004

Balasubbramanian, K. and Nagarajan, G., Effect of Injection Pressure on Performance, Emission and Combustion

Characteristics of High Linolenic Linseed Oil Methyl Ester in a DI Diesel Engine. Renewable energy. 2009; 34:1227-1233. DOI: 10.1016/j.

**112**

## Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled with Gasoline-Biodiesel Blends

*Yanuandri Putrasari and Ocktaeck Lim*

### **Abstract**

A gasoline compression ignition (GCI) engine was proposed to be the next generation internal combustion engine for gasoline. The effect of exhaust gas recirculation (EGR) and intake boosting on combustion and emissions of GCI engine fueled with gasoline-biodiesel blends by partially premixed compression ignition (PPCI) combustions are investigated in this study. Tests were conducted on a single-cylinder direct-injection CI engine, with 5% by volume proportion of biodiesel in gasoline fuel blends. Engine control parameters (EGR rate, intake boosting rate, and various injection strategies) were adjusted to investigate their influences on combustion and emissions of this GCI engine. It is found that changes in EGR rate, intake boosting pressure and injection strategies affect on ignition delay, maximum pressure rise rate and thermal efficiency which is closely tied to HC, CO, NOx and smoke emissions, respectively.

**Keywords:** gasoline, fuel, GCI, thermal efficiency, emission

#### **1. Introduction**

The stricter limitations of vehicle emission regulation especially for compression ignition (CI) engines with petroleum diesel fuel motivates many researchers to explore the utilization of low volatile and alternative fuels for CI engines to obtain high efficiency, but produce lower emissions or so-called low temperature combustion (LTC). Due to the low volatility and short ignition delay of diesel fuel, CI engines produce high nitrogen oxides (NOx) and soot emissions. To obtain LTC combustion, which is increase engine efficiency, and improve exhaust emissions, a variety of combustion methods in CI engines have been investigated such as homogeneous charge compression ignition (HCCI), PPCI, multiple premixed compression ignition (MPCI), etc. HCCI engine is one method that potentially to achieve an advanced LTC, which possible to produce low particulate matter (PM) and NOx emissions to replace conventional diesel engine combustion [1]. However, many technical problems of HCCI engine strategy must be solved before released to the market. The maximum load limitation due to the surplus of pressure rise rate (PRR) and engine knocking phenomenon [2, 3] have been major obstacle as long as HCCI combustion influenced by fuel type and air fuel mixture quality.

Recently, several studies have shown that gasoline and some other fuels with low cetane number and higher volatility are potentially advantages for low temperature combustion in CI engine, which is popular as gasoline compression ignition engine (GCI) [4–19]. At the beginning, GCI was proposed to exploit the benefit of high volatility and long ignition delay of gasoline fuel and high ratio of CI engine to obtain both high engine efficiency and low exhaust emission [20]. Later on, the concept of GCI was improved with PPCI concept by injecting fuel in the compression stroke, then mixture stratification is formed before combustion event [21]. The PPCI concept may extend load range at the same time maintaining high thermal efficiency, low NOx and soot emissions. Even though, the pressure rise rate (PRR) is remaining too high at the high load operation. The most advanced injection strategy on GCI is MPCI which has purpose to control combustion noise by manage the injection and combustion process in an order as spray – combustion – spray – combustion [22, 23]. By using MPCI mode, the acceptable pressure rise rate can be achieved and extended load range can be obtained.

strategy, initial conditions and its combustion modes. Furthermore, based on previous study [30–32], the maintaining of high efficiency and emission reduction of GCI engine fueled with gasoline-biodiesel blends still challenging and need more to be optimized especially for its NOx and soot emissions. Since the auto-ignition sensitivity of gasoline fuel is influenced by several factors such as in-cylinder equivalence ratio, intake dilution, intake temperature and pressure, it is potentially to utilize EGR and boosting in GCI combustion. To overcome the load operation limit and adjusting heat release subject to engine speed by delaying the combustion phasing, EGR was used [35]. The high-load operation can be achieved without knocking by using EGR, in which increases of specific heats capacity and minimize of oxygen (O2) concentration in the chamber promotes longer ignition delay and shifted combustion phasing far away from after top dead centre [36–39]. But, surplus of EGR supply leads to the decreasing of power and resulted more CO and HC emissions. Thus, boosting was utilized to increase the operating load simultaneously encourage fuel ignition reactivity. Furthermore, adjusting the CA50 is also necessary when boosting is operated. Because, the more intake charges mass leads to the high intensity of knocking due to the higher of pressure rise [40–42]. Suitable method of EGR and boosting was proven potentially to extend CI engine 12]. Thus, it is also potential to be used in GCI engines. The basic and control mechanism of EGR with boosting on GCI gasoline-biodiesel auto-ignition should be able to explain the relatively wide ranges of operating parameters. Thus, complementary experimental works are conducted to achieve a better understanding on the combustion process and emission characteristics of GCI engine fueled with gasoline-biodiesel blends. Information about the effects of EGR and boosting on GCI engines using gasoline-biodiesel blends are essential for advancing the theory and contributions to successfully implement gasoline in CI engines and biofuel into the transportation

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled…*

*DOI: http://dx.doi.org/10.5772/intechopen.95877*

The objective of this study was to determine the effects of EGR and boosting on

The experimental study was conducted using a single-cylinder, four-stroke, direct injection, water-cooled, naturally aspirated diesel engine with 498 cm3 of displacement and four-valve SOHC. **Figure 1** shows the schematic diagram of the test engine and measurement setup. The engine specifications are listed in **Table 1**. A standalone supercharger made by Engine Tech, a Korean local company, was used to supply the intake boosting. A conventional EGR system was used, in which the line is routed directly from exhaust manifold to the intake manifold. The engine was connected to the test system which is a 57 kW Dynamometer (Elin AVL Puma MCA325MO2). A data acquisition system (Dewetron DEWE-800-CA) in combination with an encoder (Autonics E40S8–1800-3-T-24), a pressure transducer (Kistler

the combustion and emissions of a GCI engine fueled with gasoline-biodiesel blends. To obtain a clear and comprehensive analysis of the effect of various EGR and boosting rates on combustion and emissions of GCI engine the same basic energy input of injected fuels was used for comparing the various parameters. The PPCI combustion modes for the gasoline-biodiesel blend were utilized. Modification of several initial conditions, such as intake, oil, and coolant temperatures are also conducted. The combustion characteristics of cylinder pressure, heat release rate, ignition delay, and emission characteristics are analyzed accordingly as the focus of

sector.

this study.

**2. Method**

**115**

**2.1 Test system**

There are many challenges in the implementation of GCI mode in CI engines. Several techniques are used to realize the utilization of gasoline fuel in CI engines including PPCI, MPCI and other various parameters. The start of injection (SOI) timing has effects on the balance of GCI combustion based on the study by Kodavasal et al., [24]. The study indicated that, the perfect combustion of a GCI engine was earned with start of injection (SOI) timing at -30<sup>o</sup> after top dead center (ATDC) and misfired was happened at SOI advanced than -42<sup>o</sup> ATDC. More advance SOI has potentially to solve the exhaust gas deterioration by minimizing NOx and keeps the high performance of LTC. However, more knocking, ringing, HC and CO will occur due to improper combustion. The multiple-injection strategy has been known as the solution for the NOx/soot and decrease of combustion noise with maintaining low fuel consumption in CI engine. The application of this method with higher volatility fuel which are suitable for well-premixed or properly stratified mixture prior to the ignition to reduce both NOx and soot emissions while maintaining high efficiency compared to conventional diesel CI engines was conducted previously [21, 25]. Meanwhile, the double injection strategy for pure gasoline fuel in GCI engines successfully reduced MPRR and NOx levels in half of the single-injection [17]. However, the IMEP and fuel economy were decreased. Furthermore, the significant increasing of the CO and soot emissions were also happened, even though the levels less than conventional diesel CI engine.

Blending gasoline fuel with certain percent of biodiesel in CI engines is the one way to obtain the good combustion and emissions results. Biodiesel is proven to be appropriated as a substitution fuel for CI engines [11, 26]. Furthermore, biodiesel has evidently decreasing engine exhaust pollutant [27, 28], due to the high content of oxygen which important to minimize the soot development in combustion process [29]. The implementation of gasoline-biodiesel blends in GCI engine with single and double injection modes has been studied previously [30–32]. The effect of biodiesel-gasoline blends on GCI combustion using 5% and 10% biodiesel was studied by Adams [30], which focused on reducing required intake temperature and utilization of split injection. To overcome the auto ignition difficulty of gasoline fuel without modification on intake temperature in GCI engine, the authors using a high compression ratio around 19.5 and various SOI of single injection mode in the previous study [31].

Previous studies have presented detailed analysis and discussion of the combustion and emission characteristics of GCI for PPCI or MPCI modes, by fueled with gasoline-diesel blends or gasoline biodiesel blends using direct injection GCI concept [4, 7–10, 12, 30–34]. However, the combustion and emission characteristics of CI engines are also influenced by various other factors, such as fuel injection

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled… DOI: http://dx.doi.org/10.5772/intechopen.95877*

strategy, initial conditions and its combustion modes. Furthermore, based on previous study [30–32], the maintaining of high efficiency and emission reduction of GCI engine fueled with gasoline-biodiesel blends still challenging and need more to be optimized especially for its NOx and soot emissions. Since the auto-ignition sensitivity of gasoline fuel is influenced by several factors such as in-cylinder equivalence ratio, intake dilution, intake temperature and pressure, it is potentially to utilize EGR and boosting in GCI combustion. To overcome the load operation limit and adjusting heat release subject to engine speed by delaying the combustion phasing, EGR was used [35]. The high-load operation can be achieved without knocking by using EGR, in which increases of specific heats capacity and minimize of oxygen (O2) concentration in the chamber promotes longer ignition delay and shifted combustion phasing far away from after top dead centre [36–39]. But, surplus of EGR supply leads to the decreasing of power and resulted more CO and HC emissions. Thus, boosting was utilized to increase the operating load simultaneously encourage fuel ignition reactivity. Furthermore, adjusting the CA50 is also necessary when boosting is operated. Because, the more intake charges mass leads to the high intensity of knocking due to the higher of pressure rise [40–42]. Suitable method of EGR and boosting was proven potentially to extend CI engine 12]. Thus, it is also potential to be used in GCI engines. The basic and control mechanism of EGR with boosting on GCI gasoline-biodiesel auto-ignition should be able to explain the relatively wide ranges of operating parameters. Thus, complementary experimental works are conducted to achieve a better understanding on the combustion process and emission characteristics of GCI engine fueled with gasoline-biodiesel blends. Information about the effects of EGR and boosting on GCI engines using gasoline-biodiesel blends are essential for advancing the theory and contributions to successfully implement gasoline in CI engines and biofuel into the transportation sector.

The objective of this study was to determine the effects of EGR and boosting on the combustion and emissions of a GCI engine fueled with gasoline-biodiesel blends. To obtain a clear and comprehensive analysis of the effect of various EGR and boosting rates on combustion and emissions of GCI engine the same basic energy input of injected fuels was used for comparing the various parameters. The PPCI combustion modes for the gasoline-biodiesel blend were utilized. Modification of several initial conditions, such as intake, oil, and coolant temperatures are also conducted. The combustion characteristics of cylinder pressure, heat release rate, ignition delay, and emission characteristics are analyzed accordingly as the focus of this study.

#### **2. Method**

Recently, several studies have shown that gasoline and some other fuels with low cetane number and higher volatility are potentially advantages for low temperature combustion in CI engine, which is popular as gasoline compression ignition engine (GCI) [4–19]. At the beginning, GCI was proposed to exploit the benefit of high volatility and long ignition delay of gasoline fuel and high ratio of CI engine to obtain both high engine efficiency and low exhaust emission [20]. Later on, the concept of GCI was improved with PPCI concept by injecting fuel in the compression stroke, then mixture stratification is formed before combustion event [21]. The PPCI concept may extend load range at the same time maintaining high thermal efficiency, low NOx and soot emissions. Even though, the pressure rise rate (PRR) is remaining too high at the high load operation. The most advanced injection strategy on GCI is MPCI which has purpose to control combustion noise by manage the injection and combustion process in an order as spray – combustion – spray – combustion [22, 23]. By using MPCI mode, the acceptable pressure rise rate can be

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

There are many challenges in the implementation of GCI mode in CI engines. Several techniques are used to realize the utilization of gasoline fuel in CI engines including PPCI, MPCI and other various parameters. The start of injection (SOI) timing has effects on the balance of GCI combustion based on the study by Kodavasal et al., [24]. The study indicated that, the perfect combustion of a GCI engine was earned with start of injection (SOI) timing at -30<sup>o</sup> after top dead center (ATDC) and misfired was happened at SOI advanced than -42<sup>o</sup> ATDC. More advance SOI has potentially to solve the exhaust gas deterioration by minimizing NOx and keeps the high performance of LTC. However, more knocking, ringing, HC and CO will occur due to improper combustion. The multiple-injection strategy has been known as the solution for the NOx/soot and decrease of combustion noise with maintaining low fuel consumption in CI engine. The application of this method with higher volatility fuel which are suitable for well-premixed or properly stratified mixture prior to the ignition to reduce both NOx and soot emissions while maintaining high efficiency compared to conventional diesel CI engines was conducted previously [21, 25]. Meanwhile, the double injection strategy for pure gasoline fuel in GCI engines successfully reduced MPRR and NOx levels in half of the single-injection [17]. However, the IMEP and fuel economy were decreased. Furthermore, the significant increasing of the CO and soot emissions were also happened, even though the levels less than conventional diesel CI engine.

Blending gasoline fuel with certain percent of biodiesel in CI engines is the one way to obtain the good combustion and emissions results. Biodiesel is proven to be appropriated as a substitution fuel for CI engines [11, 26]. Furthermore, biodiesel has evidently decreasing engine exhaust pollutant [27, 28], due to the high content of oxygen which important to minimize the soot development in combustion process [29]. The implementation of gasoline-biodiesel blends in GCI engine with single and double injection modes has been studied previously [30–32]. The effect of biodiesel-gasoline blends on GCI combustion using 5% and 10% biodiesel was studied by Adams [30], which focused on reducing required intake temperature and utilization of split injection. To overcome the auto ignition difficulty of gasoline fuel without modification on intake temperature in GCI engine, the authors using a high compression ratio around 19.5 and various SOI of single injection mode in the

Previous studies have presented detailed analysis and discussion of the combustion and emission characteristics of GCI for PPCI or MPCI modes, by fueled with gasoline-diesel blends or gasoline biodiesel blends using direct injection GCI concept [4, 7–10, 12, 30–34]. However, the combustion and emission characteristics of CI engines are also influenced by various other factors, such as fuel injection

achieved and extended load range can be obtained.

previous study [31].

**114**

#### **2.1 Test system**

The experimental study was conducted using a single-cylinder, four-stroke, direct injection, water-cooled, naturally aspirated diesel engine with 498 cm3 of displacement and four-valve SOHC. **Figure 1** shows the schematic diagram of the test engine and measurement setup. The engine specifications are listed in **Table 1**. A standalone supercharger made by Engine Tech, a Korean local company, was used to supply the intake boosting. A conventional EGR system was used, in which the line is routed directly from exhaust manifold to the intake manifold. The engine was connected to the test system which is a 57 kW Dynamometer (Elin AVL Puma MCA325MO2). A data acquisition system (Dewetron DEWE-800-CA) in combination with an encoder (Autonics E40S8–1800-3-T-24), a pressure transducer (Kistler

**2.2 Test fuels preparation**

*DOI: http://dx.doi.org/10.5772/intechopen.95877*

**2.3 Operating conditions**

Two fuels, which are diesel and gasoline-biodiesel blend were used in this study. The baseline fuels utilized in this study were commercial gasoline (GB00), neat diesel (D100) and pure soya bean biodiesel (B100). The chemical composition information of soya bean vegetable oil is given in **Table 2**. Biodiesel (5% by volume) and gasoline were blended and labeled as GB05. To maintain the homogeneity, the mixing process was conducted about 10 minutes then immediately used for experiment. The physical properties of baseline fuels and GB05 are given in **Table 3**.

The engine was operated at stable condition with fixed 1200 rpm. An injection

adopted and set for PPCI combustion mode. The total energy input of injected fuel was set at around 26 mg/cycle. The initial parameters of intake temperature, oil temperature, and coolant temperature were maintained at 85°C, 75°C, and 65°C, respectively. The reason why the intake temperature was maintained at 85°C is to promote the autoignition of the fuel easier. As already known that GB05 more likely as a pure gasoline which low autoignition characteristics. The homogeneous hot

**Fatty Acid System Name Structure Formula<sup>a</sup> Composition (wt %)**

**Test Item Unit Test Method Gasoline GB05 B100 D100** Heating Value MJ/kg ASTM D240:2009 45.86 45.32 39.79 45.93

Lubricity mm ISO 12156-1:2012 548 290 189 238

Density (15 °C) kg/m3 ISO 12185:2003 712.7 722.3 882.3 826.3

/s ISO 3104:2008 0.735 — 4.229 2.798

C ISO 3015:2008 �57 �37 3 �5

C ASTM D6749:2002 �57 �57 1 �9

Myristic Tetradecanoic 14:0 C14H28O2 0 Palmitic Hexadecanoic 16:0 C16H32O2 12 Stearic Octadecanoic 18:0 C18H36O2 3 Arachidic Eicosanoic 20:0 C20H40O2 0 Behenic Docosanoic 22:0 C22H44O2 0 Lignoceric Tetracosanoic 24:0 C24H48O2 0 Oleic cis-9-Octadecenoic 18:1 C18H34O2 23 Linoleic cis-9,cis-12-Octadecadienoic 18:2 C18H32O2 55 Linolenic cis-9,cis-12, cis-15-Octadecatrienoic 18:3 C18H30O2 6 Erucic cis-13-Docosenoic 22:1 C22H42O2 0

*xx:y indicates xx carbons in the fatty acid chain with y double bonds.*

*Chemical composition of soya bean vegetable oil.*

Kinematic Viscosity (40°C) mm<sup>2</sup>

Cloud Point <sup>o</sup>

Pour Point <sup>o</sup>

*Physical properties of the fuels.*

**Table 2.**

**Table 3.**

**117**

CA BTDC was

pressure of 70 MPa was used for PPCI. Single injection timing at 40 <sup>o</sup>

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled…*

#### **Figure 1.**

*Schematic diagram of the engine and measurement system setup.*


#### **Table 1.** *Engine specifications.*

6056A), and an amplifier (Kistler 5018) were used to obtain the combustion data. The fuel was injected to the combustion chamber using a Bosch seven-hole injector. A multi-stage injection engine controller (*Zenobalti:* ZB-8035, ZB-5100, ZB-100) was used to adjust the combustion strategy. Meanwhile, the temperatures of air intake, engine coolant, and lubricant oil were controlled by using separate temperature controller. The exhaust emissions, including unburned total hydrocarbon, carbon monoxide, and NOx were measured using a Horiba MEXA-7100DEGR. AVL 415 smoke meter was used for soot emission measurement. Some thermocouple K and RTD types were installed on the certain part of the engine to measure including the intake, oil, coolant and exhaust temperature.

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled… DOI: http://dx.doi.org/10.5772/intechopen.95877*

#### **2.2 Test fuels preparation**

Two fuels, which are diesel and gasoline-biodiesel blend were used in this study. The baseline fuels utilized in this study were commercial gasoline (GB00), neat diesel (D100) and pure soya bean biodiesel (B100). The chemical composition information of soya bean vegetable oil is given in **Table 2**. Biodiesel (5% by volume) and gasoline were blended and labeled as GB05. To maintain the homogeneity, the mixing process was conducted about 10 minutes then immediately used for experiment. The physical properties of baseline fuels and GB05 are given in **Table 3**.

#### **2.3 Operating conditions**

The engine was operated at stable condition with fixed 1200 rpm. An injection pressure of 70 MPa was used for PPCI. Single injection timing at 40 <sup>o</sup> CA BTDC was adopted and set for PPCI combustion mode. The total energy input of injected fuel was set at around 26 mg/cycle. The initial parameters of intake temperature, oil temperature, and coolant temperature were maintained at 85°C, 75°C, and 65°C, respectively. The reason why the intake temperature was maintained at 85°C is to promote the autoignition of the fuel easier. As already known that GB05 more likely as a pure gasoline which low autoignition characteristics. The homogeneous hot


#### **Table 2.**

6056A), and an amplifier (Kistler 5018) were used to obtain the combustion data. The fuel was injected to the combustion chamber using a Bosch seven-hole injector. A multi-stage injection engine controller (*Zenobalti:* ZB-8035, ZB-5100, ZB-100) was used to adjust the combustion strategy. Meanwhile, the temperatures of air intake, engine coolant, and lubricant oil were controlled by using separate temperature controller. The exhaust emissions, including unburned total hydrocarbon, carbon monoxide, and NOx were measured using a Horiba MEXA-7100DEGR. AVL 415 smoke meter was used for soot emission measurement. Some thermocouple K and RTD types were installed on the certain part of the engine to measure including

**Engine Parameters Value** Displacement 498 cm<sup>3</sup> Bore 83 mm Stroke 92 mm Compression Ratio 19.5 Con. Rod Length 145.8 mm Crank Radius 43.74 mm Valve System 4-valve SOHC Fuel System Electronic Common Rail

the intake, oil, coolant and exhaust temperature.

*Schematic diagram of the engine and measurement system setup.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**Figure 1.**

**Table 1.**

**116**

*Engine specifications.*

*Chemical composition of soya bean vegetable oil.*


### **Table 3.**

*Physical properties of the fuels.*

EGR and air mixture were applied in this study with 0%, 20% and 50% of flow rates by using a pair of gate valve. The EGR ratio was calculated using Eq. 1 as follows.

$$EGR\% = \frac{m\_E}{m\_E + m\_i} \times 100\% \tag{1}$$

Furthermore, the emission of CO, HC, NOx and particulate matter (smoke)

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled…*

The main purpose of this experiment is to improve the efficiency and emission characteristics of CI engine fueled with gasoline-biodiesel blends using GCI mode. To replace the utilization of diesel fuel with gasoline fuel in CI engine due to high demand of diesel fuel in the market, thus a small amount of biodiesel (5%) was added as the lubricity improver to overcome the wear problems in the fuel system. The performance results were compared with only pure diesel fuel, because the basic of the engine is diesel engine. The engine was run on single injections mode (PPCI) at 1200 rpm with various EGR rate (0%, 20% and 50%) and intake boosting 0.1 to 0.12 MPa to investigate the effect of EGR and boosting on performance, combustion and emissions. The performance, combustion and emissions characteristics data was analyzed and presented graphically for in-cylinder pressure, temperature, HRR, ignition delay, MPRR, PPRR, IMEP, thermal efficiency and its

The total fuel consumption per cycle in PPCI mode is maintained at 26 mg per

**Figure 2** shows the in-cylinder pressure, temperature and HRR of PPCI mode at various EGR rate for 0%, 20% and 50%, and fixed intake boosting 0.1 MPa. It can be seen from the figure that CI engine fueled with diesel fuel reveal the decreasing incylinder pressure when the EGR rate is increase. Similar with the diesel fuel, gasoline-biodiesel blend also indicates the same trend when EGR rate increase the in-cylinder pressure decrease. The in-cylinder temperature for diesel fuel decreasing as the trend of in-cylinder pressure when EGR rate increase. However, the incylinder temperature trends of gasoline-biodiesel blend show that EGR 50% lead to the highest value among the others EGR rates. Observing at heat release rates curves, it is seen that the heat release process of both diesel and gasoline-biodiesel blends fuels show a marked two-stage ignition. The first stage ignition of diesel fuel consistently higher than 20 J/deg., even though all of the curves reveal decreasing trends for various increasing EGR rate. Meanwhile, the first stage ignitions from gasoline-biodiesel blends are very low for all various EGR rates, and it is almost very difficult to be recognized. The highest peak of heat release rate can be obtained from gasoline-biodiesel blends with 50% EGR rate. The highest peak of heat release rate can be used to determine that the excessive pressure rise rate is happened. The excessive of PRR means that the combustion is not stable or some time when in the high load condition, the rapid pressure rise rate can result in heavy knocking

**Figure 3** shows the effect of EGR on ignition delay when engine operated using PPCI mode. The higher EGR rate results the longer ignition delay for both of diesel and gasoline-biodiesel blends. However, it can be observed that gasoline-biodiesel blends lead to the much longer ignition delay compared to diesel fuel in every EGR

conditions i.e. air intake, engine coolant and engine oil temperatures were set at 358 K, 338 K and 348 K, respectively. Meanwhile, the fixed intake pressure 0.1 MPa and various EGR rates for 0%, 20% and 50% were used to characterized the effect of EGR and PPCI injection strategy on combustion and emissions of GCI engine

CA BTDC. The others engine operating

were also discussed and analyzed in detail.

*DOI: http://dx.doi.org/10.5772/intechopen.95877*

**3. Experimental results and discussion**

emissions including HC, CO, NOx and smoke opacity.

**3.1 Effect of EGR and PPCI injection strategy**

cycle and single injection timing at 40 <sup>o</sup>

fueled with gasoline-biodiesel blends.

operation.

**119**

where mE and mi are the mass of EGR and intake fresh air, respectively.

The air boosting were set at 0.1 and 0.12 MPa in the intake manifold. The more detail engine operating parameters and injection strategies are presented in **Tables 4** and **5**, respectively. The data of 100 consecutive cycles such as in-cylinder pressure was recorded for combustion analysis.

The analysis and discussion was performed on several engine parameters such as in cylinder heat release rate, temperature, peak of pressure rise rate, IMEP, COV of IMEP, knocking/ringing intensity, thermal efficiency and combustion efficiency. Rate of heat release was calculated using Eq. 2.

$$\frac{dQ}{d\theta} = \frac{1}{\chi - 1} V \frac{dp}{d\theta} + \frac{\chi}{\chi - 1} p \frac{dV}{d\theta} \tag{2}$$

Where, γ is the specific heat ratio, V is the instantaneous cylinder volume, and p is the cylinder pressure. The normal and suitable value of γ for a CI engine is 1.3. The in-cylinder pressure and volume data were used to calculate the in-cylinder temperature using ideal gas law, as shown in Eq. 3.

$$T = \frac{p.V}{n.R} \tag{3}$$

where p is for pressure, V for volume, n is the amount of substance, and R is the gas constant.


**Table 4.**

*Operating parameters.*


**Table 5.** *Injection strategies.*

Furthermore, the emission of CO, HC, NOx and particulate matter (smoke) were also discussed and analyzed in detail.

#### **3. Experimental results and discussion**

EGR and air mixture were applied in this study with 0%, 20% and 50% of flow rates by using a pair of gate valve. The EGR ratio was calculated using Eq. 1 as follows.

*mE* þ *mi*

**Tables 4** and **5**, respectively. The data of 100 consecutive cycles such as in-cylinder

*<sup>V</sup> dp <sup>d</sup><sup>θ</sup>* <sup>þ</sup> *<sup>γ</sup> γ* � 1 *p dV*

Where, γ is the specific heat ratio, V is the instantaneous cylinder volume, and p is the cylinder pressure. The normal and suitable value of γ for a CI engine is 1.3. The in-cylinder pressure and volume data were used to calculate the in-cylinder

*<sup>T</sup>* <sup>¼</sup> *<sup>p</sup>:<sup>V</sup>*

**Parameter Diesel/GCI** Speed (rpm) 1200 Inj. Pressure (MPa) 70 Injection strategy PPCI Inj. Quantity (mg) 26 T intake (°C) 85 T oil (°C) 75 T coolant (°C) 65 EGR (%) 0, 20 and 50 Intake boosting (MPa) 0.1 and 0.12

**Combustion modes Injection strategies Injection timing and duration Injected fuel**

where p is for pressure, V for volume, n is the amount of substance, and R is the

The analysis and discussion was performed on several engine parameters such as in cylinder heat release rate, temperature, peak of pressure rise rate, IMEP, COV of IMEP, knocking/ringing intensity, thermal efficiency and combustion efficiency.

where mE and mi are the mass of EGR and intake fresh air, respectively. The air boosting were set at 0.1 and 0.12 MPa in the intake manifold. The more

detail engine operating parameters and injection strategies are presented in

pressure was recorded for combustion analysis.

Rate of heat release was calculated using Eq. 2.

temperature using ideal gas law, as shown in Eq. 3.

PPCI Single 40 <sup>o</sup>

gas constant.

**Table 4.**

**Table 5.** *Injection strategies.*

**118**

*Operating parameters.*

*dQ <sup>d</sup><sup>θ</sup>* <sup>¼</sup> <sup>1</sup> *γ* � 1 � 100% (1)

*<sup>d</sup><sup>θ</sup>* (2)

**D100 GB05**

CA BTDC (1000 μs) (26 mg)

*<sup>n</sup>:<sup>R</sup>* (3)

*EGR*% <sup>¼</sup> *mE*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

The main purpose of this experiment is to improve the efficiency and emission characteristics of CI engine fueled with gasoline-biodiesel blends using GCI mode. To replace the utilization of diesel fuel with gasoline fuel in CI engine due to high demand of diesel fuel in the market, thus a small amount of biodiesel (5%) was added as the lubricity improver to overcome the wear problems in the fuel system. The performance results were compared with only pure diesel fuel, because the basic of the engine is diesel engine. The engine was run on single injections mode (PPCI) at 1200 rpm with various EGR rate (0%, 20% and 50%) and intake boosting 0.1 to 0.12 MPa to investigate the effect of EGR and boosting on performance, combustion and emissions. The performance, combustion and emissions characteristics data was analyzed and presented graphically for in-cylinder pressure, temperature, HRR, ignition delay, MPRR, PPRR, IMEP, thermal efficiency and its emissions including HC, CO, NOx and smoke opacity.

#### **3.1 Effect of EGR and PPCI injection strategy**

The total fuel consumption per cycle in PPCI mode is maintained at 26 mg per cycle and single injection timing at 40 <sup>o</sup> CA BTDC. The others engine operating conditions i.e. air intake, engine coolant and engine oil temperatures were set at 358 K, 338 K and 348 K, respectively. Meanwhile, the fixed intake pressure 0.1 MPa and various EGR rates for 0%, 20% and 50% were used to characterized the effect of EGR and PPCI injection strategy on combustion and emissions of GCI engine fueled with gasoline-biodiesel blends.

**Figure 2** shows the in-cylinder pressure, temperature and HRR of PPCI mode at various EGR rate for 0%, 20% and 50%, and fixed intake boosting 0.1 MPa. It can be seen from the figure that CI engine fueled with diesel fuel reveal the decreasing incylinder pressure when the EGR rate is increase. Similar with the diesel fuel, gasoline-biodiesel blend also indicates the same trend when EGR rate increase the in-cylinder pressure decrease. The in-cylinder temperature for diesel fuel decreasing as the trend of in-cylinder pressure when EGR rate increase. However, the incylinder temperature trends of gasoline-biodiesel blend show that EGR 50% lead to the highest value among the others EGR rates. Observing at heat release rates curves, it is seen that the heat release process of both diesel and gasoline-biodiesel blends fuels show a marked two-stage ignition. The first stage ignition of diesel fuel consistently higher than 20 J/deg., even though all of the curves reveal decreasing trends for various increasing EGR rate. Meanwhile, the first stage ignitions from gasoline-biodiesel blends are very low for all various EGR rates, and it is almost very difficult to be recognized. The highest peak of heat release rate can be obtained from gasoline-biodiesel blends with 50% EGR rate. The highest peak of heat release rate can be used to determine that the excessive pressure rise rate is happened. The excessive of PRR means that the combustion is not stable or some time when in the high load condition, the rapid pressure rise rate can result in heavy knocking operation.

**Figure 3** shows the effect of EGR on ignition delay when engine operated using PPCI mode. The higher EGR rate results the longer ignition delay for both of diesel and gasoline-biodiesel blends. However, it can be observed that gasoline-biodiesel blends lead to the much longer ignition delay compared to diesel fuel in every EGR

**Figure 2.** *Effect of EGR on cylinder pressure, temperature and HRR of PPCI mode.*

rate variations. This condition is the advantage of gasoline fuel, which is longer ignition delay due to high volatile and low cetane number, thus there is a possibility of complete mixing period before combustion occurred. However, the longer ignition delay caused the shifted of maximum in-cylinder pressure far away form TDC which can reduce the performance of the engine. To overcome this condition the earlier injection timing can be applied among many other solutions.

indicated thermal efficiency also can be calculated by using it derivative that is indicated power/work. The effect of various EGR rates on indicated thermal efficiency of GCI engine using PPCI strategy can be seen in **Figure 6**. It can be seen that by increasing EGR rate the value of indicated thermal efficiencies are decreased for both of diesel and gasoline-biodiesel blends. The 50% EGR rate for diesel fuel lead to a little increasing value of indicated thermal efficiency is compared with 20% of EGR rate. However, it caused the significant drop value of indicated thermal effi-

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled…*

*DOI: http://dx.doi.org/10.5772/intechopen.95877*

Effect of EGR rates on CO emission of GCI engine using PPCI mode can be observed on **Figure 7**. All variation of EGR rates showed that CO emission of gasoline-biodiesel blends are lower than diesel fuel due to the volatile properties of gasoline and higher oxygen content of biodiesel, which make more complete mixing and produce more perfect combustion. However, in general, the increasing of EGR rates caused no different of CO emission for both diesel and gasoline-biodiesel blends fuels. A little decreasing value of CO emission was only happened on GCI engine fueled with gasoline-biodiesel blends when running on 50% EGR rate. **Figure 8** shows the effect of various EGR rate on HC emission of GCI engine running on PPCI strategy. As like the trend of CO emission, HC emission of GCI engine fueled with gasoline-biodiesel blends was also showed a lower value compared to diesel fuel. This condition can be explained also due to the properties of gasoline fuel and the oxygen content of biodiesel. The 20% of EGR rate value gives the lowest effect of HC emission both for diesel and for gasoline-biodiesel blends. Therefore, it is assumed in the PPCI mode the 20% of EGR rate as an optimum

The NOx emission and its effect by using various EGR rate on GCI engine using PPCI mode can be seen in **Figure 9**. Normally, the increasing of EGR rates will lead to the lower NOx emission. However, in this case, for diesel fuel, the 20% of EGR

ciency in case of gasoline-biodiesel blends fuel.

*Effect of EGR on ignition delay of PPCI mode.*

**Figure 3.**

value to obtain lowest HC emission.

**121**

**Figure 4(a)** shows the effect of EGR on maximum of in-cylinder pressure and (b) peak of pressure rise rate of PPCI strategy. The higher of EGR rate generate the lower in-cylinder pressure maximum and lower the maximum pressure rise rate for both diesel fuel and gasoline-biodiesel blends. Similarly, the increasing of EGR rates also reducing the maximum of pressure rise rate for both diesel fuel and gasolinebiodiesel blends. This condition happened due to the slowdown of combustion process. One of the reasons when utilizing EGR to slowing down of combustion process is the concentration of O2 is lowered and the concentrations of CO2 and H2O unintentionally increased. Therefore, this slows down the reactions in the oxidizing direction and speeds up the reactions of reduction process direction.

The effect of EGR on IMEP of GCI engine using PPCI strategy is presented in **Figure 5**. The increasing of EGR rates does not give any effect on IMEP of GCI engine fueled with diesel fuel. However, the 50% EGR rate results the highest IMEP value for gasoline-biodiesel blends, even much higher if compared with diesel fuel that is almost 1.0 MPa. Related to the IMEP value, the engine efficiencies especially *Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled… DOI: http://dx.doi.org/10.5772/intechopen.95877*

**Figure 3.** *Effect of EGR on ignition delay of PPCI mode.*

indicated thermal efficiency also can be calculated by using it derivative that is indicated power/work. The effect of various EGR rates on indicated thermal efficiency of GCI engine using PPCI strategy can be seen in **Figure 6**. It can be seen that by increasing EGR rate the value of indicated thermal efficiencies are decreased for both of diesel and gasoline-biodiesel blends. The 50% EGR rate for diesel fuel lead to a little increasing value of indicated thermal efficiency is compared with 20% of EGR rate. However, it caused the significant drop value of indicated thermal efficiency in case of gasoline-biodiesel blends fuel.

Effect of EGR rates on CO emission of GCI engine using PPCI mode can be observed on **Figure 7**. All variation of EGR rates showed that CO emission of gasoline-biodiesel blends are lower than diesel fuel due to the volatile properties of gasoline and higher oxygen content of biodiesel, which make more complete mixing and produce more perfect combustion. However, in general, the increasing of EGR rates caused no different of CO emission for both diesel and gasoline-biodiesel blends fuels. A little decreasing value of CO emission was only happened on GCI engine fueled with gasoline-biodiesel blends when running on 50% EGR rate.

**Figure 8** shows the effect of various EGR rate on HC emission of GCI engine running on PPCI strategy. As like the trend of CO emission, HC emission of GCI engine fueled with gasoline-biodiesel blends was also showed a lower value compared to diesel fuel. This condition can be explained also due to the properties of gasoline fuel and the oxygen content of biodiesel. The 20% of EGR rate value gives the lowest effect of HC emission both for diesel and for gasoline-biodiesel blends. Therefore, it is assumed in the PPCI mode the 20% of EGR rate as an optimum value to obtain lowest HC emission.

The NOx emission and its effect by using various EGR rate on GCI engine using PPCI mode can be seen in **Figure 9**. Normally, the increasing of EGR rates will lead to the lower NOx emission. However, in this case, for diesel fuel, the 20% of EGR

rate variations. This condition is the advantage of gasoline fuel, which is longer ignition delay due to high volatile and low cetane number, thus there is a possibility of complete mixing period before combustion occurred. However, the longer ignition delay caused the shifted of maximum in-cylinder pressure far away form TDC which can reduce the performance of the engine. To overcome this condition the

**Figure 4(a)** shows the effect of EGR on maximum of in-cylinder pressure and (b) peak of pressure rise rate of PPCI strategy. The higher of EGR rate generate the lower in-cylinder pressure maximum and lower the maximum pressure rise rate for both diesel fuel and gasoline-biodiesel blends. Similarly, the increasing of EGR rates also reducing the maximum of pressure rise rate for both diesel fuel and gasolinebiodiesel blends. This condition happened due to the slowdown of combustion process. One of the reasons when utilizing EGR to slowing down of combustion process is the concentration of O2 is lowered and the concentrations of CO2 and H2O unintentionally increased. Therefore, this slows down the reactions in the oxidizing

The effect of EGR on IMEP of GCI engine using PPCI strategy is presented in **Figure 5**. The increasing of EGR rates does not give any effect on IMEP of GCI engine fueled with diesel fuel. However, the 50% EGR rate results the highest IMEP value for gasoline-biodiesel blends, even much higher if compared with diesel fuel that is almost 1.0 MPa. Related to the IMEP value, the engine efficiencies especially

earlier injection timing can be applied among many other solutions.

*Effect of EGR on cylinder pressure, temperature and HRR of PPCI mode.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**Figure 2.**

**120**

direction and speeds up the reactions of reduction process direction.

**Figure 4.** *Effect of EGR on (a) max pressure and (b) peak pressure rise rate of PPCI mode.*

rate gives highest NOx emission. Even though, when 50% EGR was applied the NOx emission will also decreasing. However, there are no effects of EGR rate variations on NOx emission of GCI engine fueled with gasoline-biodiesel blends. This

condition can be seen in the trend of graph that from the three EGR rate variation

The smoke emission of CI engine usually contrasts with NOx emission. When the NOx higher, the smoke will be a lower and vice versa. The effect of EGR rate variation on the smoke emission of GCI engine can be seen in **Figure 10**. Smoke emission of GCI engine fueled with diesel in the high level for all variation of EGR rate, even when the rate increased. However, the smoke emission of GCI engine using gasoline-biodiesel blends obtain its lowest value when EGR rates at 20%. It can be said that the optimums of EGR rate that can maintain lowest smoke emission

resulted almost same NOx emission value.

*Effect of EGR on indicated thermal efficiency of PPCI mode.*

**Figure 5.**

**Figure 6.**

**123**

*Effect of EGR on IMEP of PPCI mode.*

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled…*

*DOI: http://dx.doi.org/10.5772/intechopen.95877*

while lowest NOx emission is 20%.

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled… DOI: http://dx.doi.org/10.5772/intechopen.95877*

**Figure 5.** *Effect of EGR on IMEP of PPCI mode.*

**Figure 6.** *Effect of EGR on indicated thermal efficiency of PPCI mode.*

condition can be seen in the trend of graph that from the three EGR rate variation resulted almost same NOx emission value.

The smoke emission of CI engine usually contrasts with NOx emission. When the NOx higher, the smoke will be a lower and vice versa. The effect of EGR rate variation on the smoke emission of GCI engine can be seen in **Figure 10**. Smoke emission of GCI engine fueled with diesel in the high level for all variation of EGR rate, even when the rate increased. However, the smoke emission of GCI engine using gasoline-biodiesel blends obtain its lowest value when EGR rates at 20%. It can be said that the optimums of EGR rate that can maintain lowest smoke emission while lowest NOx emission is 20%.

rate gives highest NOx emission. Even though, when 50% EGR was applied the NOx emission will also decreasing. However, there are no effects of EGR rate variations

on NOx emission of GCI engine fueled with gasoline-biodiesel blends. This

*Effect of EGR on (a) max pressure and (b) peak pressure rise rate of PPCI mode.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**Figure 4.**

**122**

**Figure 7.** *Effect of EGR on CO emission of PPCI mode.*

**Figure 8.** *Effect of EGR on HC emission of PPCI mode.*

#### **3.2 Effect of intake boosting and PPCI injection strategy**

To understand the effects of intake boosting on GCI engine fueled with gasolinebiodiesel blends on PPCI mode in a simple and easy way, only the 20% of EGR rate was chosen an explained in this study. The intake boosting was set at 0.1 MPa and 0.12 MPa. **Figure 11** shows the effect of boosting on in-cylinder pressure, temperature, and heat release rate of GCI engine fueled with gasoline-biodiesel blends when running on PPCI strategy. Normally found that the increasing of intake boosting rate, increasing the in-cylinder pressure for both diesel fuel and gasoline-biodiesel blends fuel. An ambient pressure of intake boosting gives a higher in-cylinder pressure of GCI engine fueled with diesel compared to gasoline-biodiesel blends. Even, the in-cylinder of gasoline biodiesel-blends with intake boosting 0.12 MPa is lower than diesel fuel with ambient intake boosting. It was also same, that the implementation of 0.12 MPa intake-boosting leads to a higher in-cylinder pressure

of GCI engine fueled with gasoline-biodiesel blends than gasoline-biodiesel with 0.1 MPa intake boosting. To obtain in cylinder pressure of gasoline biodiesel blend at least equal to pure diesel fuel, the higher intake boosting can be applied as long as the engine material supports for high pressure condition and the real engine booster in this case turbocharger can achieves maximum desired pressure. Similar with incylinder pressure, the in-cylinder temperature curves show that the highest value is for GCI engine fueled with diesel fuel when intake boosting 0.12 MPa was applied. The lowest in-cylinder temperature, which is below 2000 K, was happened for GCI engine fueled with gasoline-biodiesel blends fuel when using ambient pressure 0.12 MPa. The HRR curves show that the highest value is for GCI engine fueled with gasoline-biodiesel fuel using 0.1 MPa intake boosting. The higher HRR value, the higher-pressure rise rate that can be determines the more unstable engine combustion. The lowest HRR value was obtained from GCI engine fueled with diesel fuel in

**Figure 9.**

**Figure 10.**

**125**

*Effect of EGR on NOx emission of PPCI mode.*

*DOI: http://dx.doi.org/10.5772/intechopen.95877*

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled…*

*Effect of EGR on smoke emission of PPCI mode.*

the ambient pressure condition, which is the most stable combustion.

The effect of intake boosting on ignition delay of GCI engine using PPCI strategy is presented in **Figure 12**. The intake boosting gives effect on the lower ignition

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled… DOI: http://dx.doi.org/10.5772/intechopen.95877*

**Figure 9.** *Effect of EGR on NOx emission of PPCI mode.*

**Figure 10.** *Effect of EGR on smoke emission of PPCI mode.*

of GCI engine fueled with gasoline-biodiesel blends than gasoline-biodiesel with 0.1 MPa intake boosting. To obtain in cylinder pressure of gasoline biodiesel blend at least equal to pure diesel fuel, the higher intake boosting can be applied as long as the engine material supports for high pressure condition and the real engine booster in this case turbocharger can achieves maximum desired pressure. Similar with incylinder pressure, the in-cylinder temperature curves show that the highest value is for GCI engine fueled with diesel fuel when intake boosting 0.12 MPa was applied. The lowest in-cylinder temperature, which is below 2000 K, was happened for GCI engine fueled with gasoline-biodiesel blends fuel when using ambient pressure 0.12 MPa. The HRR curves show that the highest value is for GCI engine fueled with gasoline-biodiesel fuel using 0.1 MPa intake boosting. The higher HRR value, the higher-pressure rise rate that can be determines the more unstable engine combustion. The lowest HRR value was obtained from GCI engine fueled with diesel fuel in the ambient pressure condition, which is the most stable combustion.

The effect of intake boosting on ignition delay of GCI engine using PPCI strategy is presented in **Figure 12**. The intake boosting gives effect on the lower ignition

**3.2 Effect of intake boosting and PPCI injection strategy**

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**Figure 7.**

**Figure 8.**

**124**

*Effect of EGR on CO emission of PPCI mode.*

*Effect of EGR on HC emission of PPCI mode.*

To understand the effects of intake boosting on GCI engine fueled with gasolinebiodiesel blends on PPCI mode in a simple and easy way, only the 20% of EGR rate was chosen an explained in this study. The intake boosting was set at 0.1 MPa and 0.12 MPa. **Figure 11** shows the effect of boosting on in-cylinder pressure, temperature, and heat release rate of GCI engine fueled with gasoline-biodiesel blends when running on PPCI strategy. Normally found that the increasing of intake boosting rate, increasing the in-cylinder pressure for both diesel fuel and gasoline-biodiesel blends fuel. An ambient pressure of intake boosting gives a higher in-cylinder pressure of GCI engine fueled with diesel compared to gasoline-biodiesel blends. Even, the in-cylinder of gasoline biodiesel-blends with intake boosting 0.12 MPa is lower than diesel fuel with ambient intake boosting. It was also same, that the implementation of 0.12 MPa intake-boosting leads to a higher in-cylinder pressure

however, too long ignition delay timing sometimes caused problem in the engine

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled…*

**Figure 13** shows the effect of various intakes boosting on maximum of incylinder pressure and its maximum pressure rise rate. A normal condition happened that the increasing intake boosting, the increasing maximum in-cylinder pressure for both diesel and gasoline-biodiesel blends fuels. However, the increasing level of maximum in-cylinder pressure of diesel fuel is much higher than gasoline-biodiesel fuel. It is suspected that intake boosting caused the mixing of air fuel in diesel fuel more optimum than gasoline-biodiesel blend. The increasing of intake boosting leads to the increasing maximum pressure rise rate of GCI engine fueled with gasoline-biodiesel blend in almost same value with diesel fuel. It is mean that the GCI engine running with intake boosting for gasoline-biodiesel blend has an almost similar stability compared with diesel fuel. However, very high-pressure rise rate

Effect of various intakes boosting on IMEP of GCI engine fueled with gasolinebiodiesel blends in PPCI strategy can be seen in **Figure 14**. The IMEP of GCI engine fueled with gasoline-biodiesel blend in ambient pressure of intake boosting is higher than when intake boosting is 0.12 MPa. The opposite condition was happened for diesel fuel, which is the IMEP value of GCI engine is higher when 0.12 MPa intake boosting was applied compared with ambient pressure. The condition for IMEP of diesel fuel as the effect of increasing the intake boosting is the normal phenomenon; however, for gasoline-biodiesel blend it is quiet special. This condition suspected by the effect of high volatile and low cetane number of gasolines, which resulted higher-pressure rise rate as shown in **Figure 13**. Fluctuate of in-cylinder pressure

The indicated thermal efficiency of GCI engine using PPCI strategy affected by various intake boosting is presented in **Figure 15**. The indicated thermal efficiency of GCI engine fueled with diesel fuel increased due to the increasing of intake boosting. Similarly, for GCI engine fueled with gasoline-biodiesel blend, even though the IMEP reduced when the intake boosting increased to be 0.12 MPa. This condition, in any case, is expected in the GCI engine fueled with gasoline-biodiesel blend. Furthermore, both for ambient and 0.12 MPa intake boosting showed that

may lead to the unstable combustion and resulted the lower IMEP value.

emission and efficiency.

*Effect of boosting on ignition delay of PPCI mode.*

*DOI: http://dx.doi.org/10.5772/intechopen.95877*

**Figure 12.**

**127**

indicated that the engine in unstable condition.

**Figure 11.** *Effect of boosting on cylinder pressure, temperature, and heat release rate of PPCI mode.*

delay for both diesel and gasoline-biodiesel blend fuel. The ambient pressure of intake boosting resulted ignition delay timing for diesel fuel at around 25 <sup>o</sup> CA BTDC, then the 0.12 MPa intake boosting lead to the slightly earlier of ignition delay timing at around 27 <sup>o</sup> CA BTDC. Similar trend happened on gasoline-biodiesel fuel, that ambient pressure of intake boosting resulted ignition delay timing at around 11 o CA BTDC, then when 0.12 MPa intake boosting was applied the ignition delay timing also more advanced at around 2 <sup>o</sup> CA BTDC. The higher volatile and lower cetane number properties of gasoline fuel caused the longer ignition delay timing if compared with diesel fuel. However, the application of intake boosting resulted a shifting of ignition delay timing earlier. The longer ignition delay timing is possible to produce more complete mixing period of air and fuel prior to combustion,

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled… DOI: http://dx.doi.org/10.5772/intechopen.95877*

#### **Figure 12.** *Effect of boosting on ignition delay of PPCI mode.*

however, too long ignition delay timing sometimes caused problem in the engine emission and efficiency.

**Figure 13** shows the effect of various intakes boosting on maximum of incylinder pressure and its maximum pressure rise rate. A normal condition happened that the increasing intake boosting, the increasing maximum in-cylinder pressure for both diesel and gasoline-biodiesel blends fuels. However, the increasing level of maximum in-cylinder pressure of diesel fuel is much higher than gasoline-biodiesel fuel. It is suspected that intake boosting caused the mixing of air fuel in diesel fuel more optimum than gasoline-biodiesel blend. The increasing of intake boosting leads to the increasing maximum pressure rise rate of GCI engine fueled with gasoline-biodiesel blend in almost same value with diesel fuel. It is mean that the GCI engine running with intake boosting for gasoline-biodiesel blend has an almost similar stability compared with diesel fuel. However, very high-pressure rise rate indicated that the engine in unstable condition.

Effect of various intakes boosting on IMEP of GCI engine fueled with gasolinebiodiesel blends in PPCI strategy can be seen in **Figure 14**. The IMEP of GCI engine fueled with gasoline-biodiesel blend in ambient pressure of intake boosting is higher than when intake boosting is 0.12 MPa. The opposite condition was happened for diesel fuel, which is the IMEP value of GCI engine is higher when 0.12 MPa intake boosting was applied compared with ambient pressure. The condition for IMEP of diesel fuel as the effect of increasing the intake boosting is the normal phenomenon; however, for gasoline-biodiesel blend it is quiet special. This condition suspected by the effect of high volatile and low cetane number of gasolines, which resulted higher-pressure rise rate as shown in **Figure 13**. Fluctuate of in-cylinder pressure may lead to the unstable combustion and resulted the lower IMEP value.

The indicated thermal efficiency of GCI engine using PPCI strategy affected by various intake boosting is presented in **Figure 15**. The indicated thermal efficiency of GCI engine fueled with diesel fuel increased due to the increasing of intake boosting. Similarly, for GCI engine fueled with gasoline-biodiesel blend, even though the IMEP reduced when the intake boosting increased to be 0.12 MPa. This condition, in any case, is expected in the GCI engine fueled with gasoline-biodiesel blend. Furthermore, both for ambient and 0.12 MPa intake boosting showed that

delay for both diesel and gasoline-biodiesel blend fuel. The ambient pressure of intake boosting resulted ignition delay timing for diesel fuel at around 25 <sup>o</sup>

*Effect of boosting on cylinder pressure, temperature, and heat release rate of PPCI mode.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

timing at around 27 <sup>o</sup>

timing also more advanced at around 2 <sup>o</sup>

o

**126**

**Figure 11.**

BTDC, then the 0.12 MPa intake boosting lead to the slightly earlier of ignition delay

that ambient pressure of intake boosting resulted ignition delay timing at around 11

cetane number properties of gasoline fuel caused the longer ignition delay timing if compared with diesel fuel. However, the application of intake boosting resulted a shifting of ignition delay timing earlier. The longer ignition delay timing is possible to produce more complete mixing period of air and fuel prior to combustion,

CA BTDC, then when 0.12 MPa intake boosting was applied the ignition delay

CA BTDC. Similar trend happened on gasoline-biodiesel fuel,

CA BTDC. The higher volatile and lower

CA

**Figure 13.** *Effect of boosting on (a) max pressure and (b) peak pressure rise rate of PPCI mode.*

the indicated thermal efficiency of GCI engine with diesel fuel is higher than gasoline-biodiesel blend.

**Figure 16** shows the effect of intake boosting on CO emission of GCI engine using PPCI strategy. It is already known that the utilization of gasoline-biodiesel blend in GCI engine resulted lower CO emission compared to diesel fuel. Similarly, in the single injection method of PPCI strategy also obtained the lower CO emission of GCI engine fueled with gasoline-biodiesel blend compared to diesel fuel. The increasing of intake boosting from 0.1 to 0.12 MPa in GCI engines gives effect on the decreasing of CO emission for both gasoline-biodiesel blend and diesel fuels. It is suspected due to the combination of 20% EGR and 0.12 MPa of intake boosting, which may lead to the complete combustion.

gasoline-biodiesel blend, it is obtained greatly decreasing of HC emission when the 0.12 MPa of intake boosting applied compared with 0.1 MPa. The decreasing value of HC emission in 0.12 MPa of intake boosting is almost 90% lower from the

**Figure 18** shows the effect of intake boosting on NOx emission of GCI engine with PPCI strategy. Overall, the NOx emission of GCI engine fueled with diesel is higher than GCI engine fueled with gasoline-biodiesel blend when using PPCI mode for either ambient intake pressure or increasing intake pressure at 0.12 MPa. The trend of graph shows that the increasing intake boosting also followed by increasing the NOx emission for both diesel and gasoline-biodiesel blend. It is mean that the increasing of intake boosting has opposite function with 20% EGR. In this case, by using only 20% EGR rate, the NOx emission of GCI engine fueled with gasoline-biodiesel blend is very low under 0.05 mg/kWh. However, increasing intake boosting 0.12 MPa, leads the deterioration on NOx emission to be around

ambient pressure of intake boosting.

*Effect of boosting on indicated thermal efficiency of PPCI mode.*

0.2 mg/kWh.

**129**

**Figure 14.**

**Figure 15.**

*Effect of boosting on IMEP of PPCI mode.*

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled…*

*DOI: http://dx.doi.org/10.5772/intechopen.95877*

The effect of intake boosting on HC emission of GCI engine can be observed in **Figure 17**. Similar with the trend on CO emission, the HC emission of GCI engine fueled with of GCI engine with gasoline-biodiesel blend originally is lower than diesel fuel as it can be seen in the ambient intake boosting condition. When the intake boosting increased to be 0.12 MPa HC emission of GCI engine decreased around a half value than 0.1 MPa of intake boosting. For GCI engine fueled with

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled… DOI: http://dx.doi.org/10.5772/intechopen.95877*

**Figure 14.** *Effect of boosting on IMEP of PPCI mode.*

**Figure 15.** *Effect of boosting on indicated thermal efficiency of PPCI mode.*

gasoline-biodiesel blend, it is obtained greatly decreasing of HC emission when the 0.12 MPa of intake boosting applied compared with 0.1 MPa. The decreasing value of HC emission in 0.12 MPa of intake boosting is almost 90% lower from the ambient pressure of intake boosting.

**Figure 18** shows the effect of intake boosting on NOx emission of GCI engine with PPCI strategy. Overall, the NOx emission of GCI engine fueled with diesel is higher than GCI engine fueled with gasoline-biodiesel blend when using PPCI mode for either ambient intake pressure or increasing intake pressure at 0.12 MPa. The trend of graph shows that the increasing intake boosting also followed by increasing the NOx emission for both diesel and gasoline-biodiesel blend. It is mean that the increasing of intake boosting has opposite function with 20% EGR. In this case, by using only 20% EGR rate, the NOx emission of GCI engine fueled with gasoline-biodiesel blend is very low under 0.05 mg/kWh. However, increasing intake boosting 0.12 MPa, leads the deterioration on NOx emission to be around 0.2 mg/kWh.

the indicated thermal efficiency of GCI engine with diesel fuel is higher than

*Effect of boosting on (a) max pressure and (b) peak pressure rise rate of PPCI mode.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**Figure 16** shows the effect of intake boosting on CO emission of GCI engine using PPCI strategy. It is already known that the utilization of gasoline-biodiesel blend in GCI engine resulted lower CO emission compared to diesel fuel. Similarly, in the single injection method of PPCI strategy also obtained the lower CO emission of GCI engine fueled with gasoline-biodiesel blend compared to diesel fuel. The increasing of intake boosting from 0.1 to 0.12 MPa in GCI engines gives effect on the decreasing of CO emission for both gasoline-biodiesel blend and diesel fuels. It is suspected due to the combination of 20% EGR and 0.12 MPa of intake boosting,

The effect of intake boosting on HC emission of GCI engine can be observed in **Figure 17**. Similar with the trend on CO emission, the HC emission of GCI engine fueled with of GCI engine with gasoline-biodiesel blend originally is lower than diesel fuel as it can be seen in the ambient intake boosting condition. When the intake boosting increased to be 0.12 MPa HC emission of GCI engine decreased around a half value than 0.1 MPa of intake boosting. For GCI engine fueled with

gasoline-biodiesel blend.

**Figure 13.**

**128**

which may lead to the complete combustion.

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

**Figure 16.** *Effect of boosting on CO emission of PPCI mode.*

**4. Conclusions**

**Figure 19.**

**131**

**Figure 18.**

*Effect of boosting on NOx emission of PPCI mode.*

*DOI: http://dx.doi.org/10.5772/intechopen.95877*

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled…*

*Effect of boosting on smoke emission of PPCI mode.*

drawn from this study:

The study on GCI engine was conducted in an experiment using biodiesel addition 5% into gasoline, compared to neat diesel with single injection (PPCI) strategy combined with the application of EGR and intake boosting in order to obtain high efficiency and low emission of GCI engine. The engine testing was set in the same of energy input that is injected fuel amount around 26 mg per cycle. Based on the results and comprehensive analysis, the following general conclusions may be

Increasing EGR rate the value of indicated thermal efficiencies are decreased for both of diesel and gasoline-biodiesel blends. The highest 50% EGR rate for diesel fuel leads to a little increasing value of indicated thermal efficiency is compared with 20% of EGR rate. However, it caused the significant drop value of indicated thermal efficiency in case of gasoline-biodiesel blends fuel. By using diesel fuel, the 20% of EGR rate gives highest NOx emission. Even though, when 50% EGR was applied the NOx emission will also decreasing. The utilization of EGR gives effect on the drop of NOx emission value for gasoline-biodiesel blend much lower than diesel fuel. However, there are no effects of EGR rate variations on NOx emission of GCI engine fueled with gasoline-biodiesel blends. Smoke emission of GCI engine fueled

**Figure 17.** *Effect of boosting on HC emission of PPCI mode.*

The effect of intake boosting on smoke emission of GCI engine with PPCI strategy can be seen in **Figure 19**. Smoke emission of GCI engine fueled with diesel fuel is very high almost 6 g/m<sup>3</sup> when running on PPCI mode by 20% of EGR rate and ambient pressure of intake boosting. While, in this condition smoke emission of GCI engine fueled with gasoline-biodiesel blend much lower than diesel fuel at around 1.5 g/m<sup>3</sup> . Increasing intake boosting to be 0.12 MPa makes smoke emission of GCI engine fueled with diesel fuel decrease very significant around 3 g/m<sup>3</sup> . However, the increasing of intake boosting to be 0.12 MPa for GCI engine fueled with gasoline-biodiesel caused the increasing of smoke emission, even though still lower than the emission of GCI engine fueled with diesel fuel which is to be around 2.5 g/m<sup>3</sup> . If the point of view of GCI engine fueled with gasoline-biodiesel blend running on PPCI mode focused simultaneously on NOx emission and smoke emission, then it can be stated that the optimum effort to reduce both of emission parts is by using 20% EGR rate and 0.1 MPa intake boosting.

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled… DOI: http://dx.doi.org/10.5772/intechopen.95877*

**Figure 18.**

*Effect of boosting on NOx emission of PPCI mode.*

**Figure 19.** *Effect of boosting on smoke emission of PPCI mode.*

#### **4. Conclusions**

The study on GCI engine was conducted in an experiment using biodiesel addition 5% into gasoline, compared to neat diesel with single injection (PPCI) strategy combined with the application of EGR and intake boosting in order to obtain high efficiency and low emission of GCI engine. The engine testing was set in the same of energy input that is injected fuel amount around 26 mg per cycle. Based on the results and comprehensive analysis, the following general conclusions may be drawn from this study:

Increasing EGR rate the value of indicated thermal efficiencies are decreased for both of diesel and gasoline-biodiesel blends. The highest 50% EGR rate for diesel fuel leads to a little increasing value of indicated thermal efficiency is compared with 20% of EGR rate. However, it caused the significant drop value of indicated thermal efficiency in case of gasoline-biodiesel blends fuel. By using diesel fuel, the 20% of EGR rate gives highest NOx emission. Even though, when 50% EGR was applied the NOx emission will also decreasing. The utilization of EGR gives effect on the drop of NOx emission value for gasoline-biodiesel blend much lower than diesel fuel. However, there are no effects of EGR rate variations on NOx emission of GCI engine fueled with gasoline-biodiesel blends. Smoke emission of GCI engine fueled

The effect of intake boosting on smoke emission of GCI engine with PPCI strategy can be seen in **Figure 19**. Smoke emission of GCI engine fueled with diesel fuel is very high almost 6 g/m<sup>3</sup> when running on PPCI mode by 20% of EGR rate and ambient pressure of intake boosting. While, in this condition smoke emission of GCI engine fueled with gasoline-biodiesel blend much lower than diesel fuel at

of GCI engine fueled with diesel fuel decrease very significant around 3 g/m<sup>3</sup>

is by using 20% EGR rate and 0.1 MPa intake boosting.

However, the increasing of intake boosting to be 0.12 MPa for GCI engine fueled with gasoline-biodiesel caused the increasing of smoke emission, even though still lower than the emission of GCI engine fueled with diesel fuel which is to be around

. If the point of view of GCI engine fueled with gasoline-biodiesel blend running on PPCI mode focused simultaneously on NOx emission and smoke emission, then it can be stated that the optimum effort to reduce both of emission parts

. Increasing intake boosting to be 0.12 MPa makes smoke emission

.

around 1.5 g/m<sup>3</sup>

2.5 g/m<sup>3</sup>

**130**

**Figure 17.**

**Figure 16.**

*Effect of boosting on CO emission of PPCI mode.*

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

*Effect of boosting on HC emission of PPCI mode.*

with diesel in the high level for all variation of EGR rate, even when the rate increased. However, the smoke emission of GCI engine using gasoline-biodiesel blends obtains its lowest value when EGR rates at 20%.

**References**

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2002-01-2859.

[1] Dec JE. Advanced compressionignition engines - Understanding the incylinder processes. Proc Combust Inst 2009;32 II:2727–42. doi:10.1016/j.

*DOI: http://dx.doi.org/10.5772/intechopen.95877*

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled…*

Engine Operating with In-Cylinder Gasoline and Diesel Fuel Blending. SAE

[9] Leermakers CAJ, Van den Berge B, Luijten CCM, Somers LMT, de Goey LPH, Albrecht BA. Gasoline-Diesel Dual Fuel: Effect of Injection Timing and Fuel Balance. SAE Pap 2011; 2011–01–24. doi:10.4271/2011-01-2437.

[10] Liu H, Wang Z, Wang J, He X.

characteristics and thermal ef fi ciency in diesel engines by fueling gasoline / diesel / PODEn blends. Energy 2016;97: 105–12. doi:10.1016/j.energy.2015.

[11] Bae C, Kim J. Alternative fuels for internal combustion engines. Proc Combust Inst 2017;36:3389–413. doi:

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doi:10.1016/j.fuel.2015.07.009.

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Comparative behavior of gasoline-diesel / butanol-diesel blends and injection strategy management on performance and emissions of a light duty diesel engine. Energy 2014;71:321–31. doi: 10.1016/j.energy.2014.04.065.

[8] Prikhodko VY, Curran SJ, Barone TL, Lewis S a, Storey JM, Cho K, et al. Emission Characteristics of a Diesel

[7] Iannuzzi SE, Valentino G.

The indicated thermal efficiency of GCI engine fueled with diesel fuel increased due to the increasing of intake boosting. Similarly, for GCI engine fueled with gasoline-biodiesel blend, the indicated thermal efficiency was also increased when the intake boosting increased to be 0.12 MPa. The NOx emission of GCI engine fueled with diesel is higher than GCI engine fueled with gasoline-biodiesel blend when using PPCI mode for either ambient intake pressure or increasing intake pressure at 0.12 MPa. The increasing intake boosting also followed by increasing NOx emission for both diesel and gasoline-biodiesel blend. Increasing intake boosting to be 0.12 MPa makes smoke emission of GCI engine fueled with diesel fuel decrease very significant around 3 g/m3 . However, the increasing of intake boosting to be 0.12 MPa for GCI engine fueled with gasoline-biodiesel caused the increasing of smoke emission, even though still lower than the emission of GCI engine fueled with diesel fuel.

#### **Acknowledgements**

This research was supported by the University of Ulsan, Korea. Yanuandri Putrasari acknowledges the support from the Indonesian Institute of Sciences (LIPI) and the Ministry of Research and Technology/National Research and Innovation Agency of Republic Indonesia (RISTEK-BRIN).

#### **Author contribution**

All authors contributed equally as the main contributor of this chapter. All authors read and approved the final version of this chapter.

#### **Conflict of interest**

The authors declare no conflict of interest.

#### **Author details**

Yanuandri Putrasari<sup>1</sup> and Ocktaeck Lim<sup>2</sup> \*

1 Research Centre for Electrical Power and Mechatronics – Indonesian Institute of Sciences (LIPI), Bandung, Indonesia

2 School of Mechanical Engineering, University of Ulsan, Ulsan, South Korea

\*Address all correspondence to: otlim@ulsan.ac.kr

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

*Combustion and Emissions of Gasoline Compression Ignition Engine Fuelled… DOI: http://dx.doi.org/10.5772/intechopen.95877*

#### **References**

with diesel in the high level for all variation of EGR rate, even when the rate increased. However, the smoke emission of GCI engine using gasoline-biodiesel

due to the increasing of intake boosting. Similarly, for GCI engine fueled with gasoline-biodiesel blend, the indicated thermal efficiency was also increased when the intake boosting increased to be 0.12 MPa. The NOx emission of GCI engine fueled with diesel is higher than GCI engine fueled with gasoline-biodiesel blend when using PPCI mode for either ambient intake pressure or increasing intake pressure at 0.12 MPa. The increasing intake boosting also followed by increasing NOx emission for both diesel and gasoline-biodiesel blend. Increasing intake boosting to be 0.12 MPa makes smoke emission of GCI engine fueled with diesel

boosting to be 0.12 MPa for GCI engine fueled with gasoline-biodiesel caused the increasing of smoke emission, even though still lower than the emission of GCI

This research was supported by the University of Ulsan, Korea. Yanuandri Putrasari acknowledges the support from the Indonesian Institute of Sciences (LIPI) and the Ministry of Research and Technology/National Research and Inno-

All authors contributed equally as the main contributor of this chapter. All

\*

1 Research Centre for Electrical Power and Mechatronics – Indonesian Institute of

© 2021 The Author(s). Licensee IntechOpen. This chapter is distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/ by/3.0), which permits unrestricted use, distribution, and reproduction in any medium,

2 School of Mechanical Engineering, University of Ulsan, Ulsan, South Korea

The indicated thermal efficiency of GCI engine fueled with diesel fuel increased

. However, the increasing of intake

blends obtains its lowest value when EGR rates at 20%.

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

fuel decrease very significant around 3 g/m3

vation Agency of Republic Indonesia (RISTEK-BRIN).

authors read and approved the final version of this chapter.

The authors declare no conflict of interest.

\*Address all correspondence to: otlim@ulsan.ac.kr

Yanuandri Putrasari<sup>1</sup> and Ocktaeck Lim<sup>2</sup>

provided the original work is properly cited.

Sciences (LIPI), Bandung, Indonesia

engine fueled with diesel fuel.

**Acknowledgements**

**Author contribution**

**Conflict of interest**

**Author details**

**132**

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[16] Yang B, Yao M, Zheng Z, Yue L. Experimental Investigation of Injection

*Internal Combustion Engine Technology and Applications of Biodiesel Fuel*

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[24] Kodavasal J, Kolodziej CP, Ciatti SA,

Sibendu S. Computational Fluid Dynamics Simulation of Gasoline Compression Ignition 2017;137:1–13. doi:

[25] Won HW, Peters N, Pitsch H, Tait N, Kalghatgi G. Partially premixed combustion of gasoline type fuels using

compression ratio in a diesel engine. SAE Tech Pap 2013;11. doi:10.4271/

[26] Tesfa B, Mishra R, Zhang C, Gu F, Ball AD. Combustion and performance characteristics of CI (compression ignition) engine running with biodiesel. Energy 2013;51:101–15. doi:10.1016/j.

[27] Cordiner S, Mulone V, Nobile M, Rocco V. Impact of biodiesel fuel on engine emissions and Aftertreatment System operation. Appl Energy 2016;

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[29] Wang Z, Li L, Wang J, Reitz RD. Effect of biodiesel saturation on soot formation in diesel engines. Fuel 2016;

[30] Adams CA, Loeper P, Krieger R, Andrie MJ, Foster DE. Effects of biodiesel-gasoline blends on gasoline direct-injection compression ignition

larger size nozzle and higher

10.1115/1.4029963.

2013-01-2539.

energy.2013.01.010.

164:972–83. doi:10.1016/j. apenergy.2015.07.001.

10.1016/j.fuel.2007.04.011.

175:240–8. doi:10.1016/j. fuel.2016.02.048.

Strategies on Low Temperature Combustion Fuelled with Gasoline in a Compression Ignition Engine 2015;2015.

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doi:10.1016/j.fuel.2017.04.107.

[18] Kodavasal J, Kolodziej CP, Ciatti SA. Effects of injection parameters , boost , and swirl ratio on gasoline compression ignition operation at idle and low-load conditions. Int J Engine Res 2016:1–13. doi:10.1177/1468087416675709.

[19] Loeper P, Ra Y, Foster D, Ghandhi J. Experimental and computational assessment of inlet swirl effects on a gasoline compression ignition (GCI) light-duty diesel engine. SAE 2014 World Congr Exhib 2014;1. doi:10.4271/

2014-01-1299.

[20] Kalghatgi GT, Risberg P,

10.4271/2006-01-3385.

10.4271/2007-01-0006.

j.fuel.2012.05.016.

**134**

Combustion and emission

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[23] Wang B, Wang Z, Shuai S, Xu H.

characteristics of Multiple Premixed

[21] Kalghatgi GT, Risberg P, Angstrom H-E. Partially Pre-Mixed Auto-Ignition of Gasoline to Attain Low Smoke and Low NOx at High Load in a Compression Ignition Engine and Comparison with a Diesel Fuel. SAE Tech Pap 2007;2007–01–00. doi:

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[33] Yang B, Li S, Zheng Z, Yao M, Cheng W. A Comparative Study on Different Dual-Fuel Combustion Modes Fuelled with Gasoline and Diesel. SAE Int 2012;694. doi:10.4271/2012-01-0694.

[34] Liu J, Shang H, Wang H, Zheng Z, Wang Q, Xue Z, et al. Investigation on partially premixed combustion fueled with gasoline and PODE blends in a multi-cylinder heavy-duty diesel engine. Fuel 2017;193:101–11. doi: 10.1016/j.fuel.2016.12.045.

[35] Cairns A, Blaxill H. The Effects of Combined Internal and External Exhaust Gas Recirculation on Gasoline Controlled Auto-Ignition 2005;2005. doi:10.4271/2005-01-0133.

[36] Zhao H, Peng Z, Williams J, Ladommatos N. Understanding the Effects of Recycled Burnt Gases on the Controlled Autoignition (CAI) Combustion in Four-Stroke Gasoline Engines 2001. doi:10.4271/2001-01- 3607.

[37] Olsson J-O, Tunestål P, Ulfvik J, Johansson B. The effect of cooled EGR on emissions and performance of a turbocharged HCCI engine. Soc Automot Eng 2003;2003:21–38. doi: 10.4271/2003-01-0743.

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### *Edited by Enhua Wang*

This book examines internal combustion engine technology and applications of biodiesel fuel. It includes seven chapters in two sections. The first section examines engine downsizing, fuel spray, and economic comparison. The second section deals with applications of biodiesel fuel in compression-ignition and spark-ignition engines. The information contained herein is useful for scientists and students looking to broaden their knowledge of internal combustion engine technologies and applications of biodiesel fuel.

Published in London, UK © 2021 IntechOpen © yucelyilmaz / iStock

Internal Combustion Engine Technology and Applications of Biodiesel Fuel

Internal Combustion Engine

Technology and Applications

of Biodiesel Fuel

*Edited by Enhua Wang*