**3. Experimental studies**

This piece of work presents the results of the tests of engine, during which distribution of fuel between the direct injection system and port injection system was changed.

For each test the constant injection and ignition timing and the stoichiometric composition of the mixture was maintained. The direct injection timing was determined in preliminary tests at 281° CA before TDC, which means direct injection of fuel during the intake stroke. Also during the preliminary test of the engine the pressure of direct fuel injection was set at 8 MPa. The injection time for both of the fuel supply systems was adjusted so as to maintain a stoichiometric mixture composition at different values of the fraction of fuel injected directly into the cylinder xDI.

## **3.1. Impact of the application of dual-injection system on the performance and fuel consumption**

On the basis of results of the above mentioned tests, the curves of torque T and brake specific fuel consumption BSFC in a function of the fraction of fuel injected directly into the cylinder xDI were obtained. Figure 12 shows the traces of torque and specific fuel consumption ap‐ proximated by parabolas obtained at the throttle opening 13% and engine rotational speed 2000 RPM.

**Figure 12.** The traces of torque and specific fuel consumption in a function of the fraction of fuel injected directly into

**Figure 11.** The scheme of the fuel system; 1 – Fuel Tank, 2 – Shutoff valve, 3 – Fuel filter, 4 – DI priming pump, 5 – Electrovalves for measurement of fuel flow in DI-circuit, 6 – Regulator of low-pressure of DI-circuit, 7 – High pressure pump, 8 – Regulator of high-pressure of DI-circuit, 9 – Engine, 10 – Direct fuel injector, 11 – Rail of the direct fuel injec‐ tors, 12 – Indirect fuel injector, 13 – Intake pipe, 14 – Rail of the indirect fuel injectors, 15 – DI pressure gauge, 16 – MPI

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For the case shown in this figure, it is seen that the maximum torque and minimum specific fuel consumption were obtained for the fraction of fuel injected directly into the cylinder xDI equal to nearly 0.4. The results obtained with this distribution of fuel between the direct

the cylinder xDI obtained for the throttle opening 13% and engine rotational speed 2000 RPM

fuel pump, 17 – Regulator of pressure of MPI-circuit, 18 – Fuel flow meter

Combustion Process in the Spark-Ignition Engine with Dual-Injection System http://dx.doi.org/10.5772/54160 63

**Figure 11.** The scheme of the fuel system; 1 – Fuel Tank, 2 – Shutoff valve, 3 – Fuel filter, 4 – DI priming pump, 5 – Electrovalves for measurement of fuel flow in DI-circuit, 6 – Regulator of low-pressure of DI-circuit, 7 – High pressure pump, 8 – Regulator of high-pressure of DI-circuit, 9 – Engine, 10 – Direct fuel injector, 11 – Rail of the direct fuel injec‐ tors, 12 – Indirect fuel injector, 13 – Intake pipe, 14 – Rail of the indirect fuel injectors, 15 – DI pressure gauge, 16 – MPI fuel pump, 17 – Regulator of pressure of MPI-circuit, 18 – Fuel flow meter

**3. Experimental studies**

9 – High pressure fuel pump, 10 – Eddy current dyno

62 Advances in Internal Combustion Engines and Fuel Technologies

into the cylinder xDI.

**consumption**

2000 RPM.

This piece of work presents the results of the tests of engine, during which distribution of fuel

**Figure 10.** The overall view of the test stand [12];1 – Engine, 2 – PC, 3 – Programmable Engine Management System, 4 – Digital oscilloscope, 5 – PC with Data Acquisition System, 6 – Throttle actuator, 7 – Fuel flow meter 8 – Gas Analyzer,

For each test the constant injection and ignition timing and the stoichiometric composition of the mixture was maintained. The direct injection timing was determined in preliminary tests at 281° CA before TDC, which means direct injection of fuel during the intake stroke. Also during the preliminary test of the engine the pressure of direct fuel injection was set at 8 MPa. The injection time for both of the fuel supply systems was adjusted so as to maintain a stoichiometric mixture composition at different values of the fraction of fuel injected directly

**3.1. Impact of the application of dual-injection system on the performance and fuel**

On the basis of results of the above mentioned tests, the curves of torque T and brake specific fuel consumption BSFC in a function of the fraction of fuel injected directly into the cylinder xDI were obtained. Figure 12 shows the traces of torque and specific fuel consumption ap‐ proximated by parabolas obtained at the throttle opening 13% and engine rotational speed

between the direct injection system and port injection system was changed.

**Figure 12.** The traces of torque and specific fuel consumption in a function of the fraction of fuel injected directly into the cylinder xDI obtained for the throttle opening 13% and engine rotational speed 2000 RPM

For the case shown in this figure, it is seen that the maximum torque and minimum specific fuel consumption were obtained for the fraction of fuel injected directly into the cylinder xDI equal to nearly 0.4. The results obtained with this distribution of fuel between the direct injection system and port injection system show significant differences especially in compar‐ ison with the test results obtained when the entire amount of fuel is injected directly into the cylinder.

Curves of torque and specific fuel consumption as a function of the fraction of fuel injected directly into the cylinder xDI obtained at 2000 RPM and throttle opening of 20% are shown in Figure 13.

**Figure 13.** The traces of torque and specific fuel consumption in a function of the fraction of fuel injected directly into the cylinder xDI obtained for the throttle opening 20% and engine rotational speed 2000 RPM

**•** For the throttle opening equal to 20% and engine speed 2000 RPM best results of specific fuel consumption and torque was observed for the ratio of fuel injected directly into the cylinder amounting to 0.62. In the described case, the mentioned operating parameters of the engine received a significant improvement in relation to the situation when the whole amount of the fuel is injected into intake channels.

Figure 14 shows charts of engine total efficiency and relative increase of the engine total efficiency ΔηDI + MPI for dual-injection operation in relation to operation with indirect fuel injection developed on the basis of the results of Figure 12 and Figure 13. The traces shown on Figure 14 are the result of parabolic approximation the points obtained by the calculations.

The engine total efficiency is determined by the formula (1). For the calculation the calorific value of petrol Wd = 44 000 kJ / kg was assumed [13].

$$
\eta\_{tot} = \frac{3.6 \cdot 10^6}{\text{BSFC} \cdot \text{W}\_d} \tag{1}
$$

observed of operation for the fraction of the fuel injected directly into the cylinder equal to 0.62. In the second situation the greatest improvement in the total efficiency of the engine with regard to the efficiency obtained with indirect fuel injection took place, when the fraction of

**Figure 14.** Engine total efficiency ηtot and relative increase of the engine total efficiency ΔηDI + MPI for dual-injection op‐

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The analysis of the results shows that, using the dual-injection system the torque generated by the engine can be improved and, what is even more importantly, the specific fuel consumption

**•** During the described above tests of the engine using gas analyzer Arcon Oliver K-4500 volumetric concentrations of individual exhaust components in the exhaust manifold were measured The concentration of carbon monoxide CO, carbon dioxide CO2, nitric oxide NO, unburned hydrocarbons HC and additionally exhaust gas temperature texh were investigat‐ ed. The total concentration of hydrocarbons in the exhaust HC was converted by the gas

In Figure 15 registered at speed 2000 RPM and at throttle opening of 13% the traces of volumetric concentrations of the above mentioned chemicals and the exhaust gas temperature

**•** The analysis of Figure 15 shows that with increase in the fraction of fuel injected directly into the cylinder the concentrations of carbon monoxide and hydrocarbons increase slightly, while the concentrations of nitrogen oxide and carbon dioxide decrease. Also the tempera‐ ture of gas leaving the engine cylinders decreased slightly. The difference between the NO concentration for injection only into the intake channel and only with direct injection into the cylinder is not high and is approximately 170 ppm. The concentration of HC for direct

were shown depending on the fraction of fuel injected directly into the cylinder.

fuel injected directly into the cylinder equals 0.39.

eration in relation to operation with indirect fuel injection

analyzer to hexane.

can be reduced. This means the improvement in the total efficiency.

**3.2. Exhaust gas composition at operation with dual-injection**

The highest increase of the total efficiency ΔηDI+MPI shown in the Figure 14 amounted to 4.58% for the first case and 2.18% in the second test point. In the first case the best efficiency was

**Figure 14.** Engine total efficiency ηtot and relative increase of the engine total efficiency ΔηDI + MPI for dual-injection op‐ eration in relation to operation with indirect fuel injection

observed of operation for the fraction of the fuel injected directly into the cylinder equal to 0.62. In the second situation the greatest improvement in the total efficiency of the engine with regard to the efficiency obtained with indirect fuel injection took place, when the fraction of fuel injected directly into the cylinder equals 0.39.

The analysis of the results shows that, using the dual-injection system the torque generated by the engine can be improved and, what is even more importantly, the specific fuel consumption can be reduced. This means the improvement in the total efficiency.

#### **3.2. Exhaust gas composition at operation with dual-injection**

injection system and port injection system show significant differences especially in compar‐ ison with the test results obtained when the entire amount of fuel is injected directly into the

Curves of torque and specific fuel consumption as a function of the fraction of fuel injected directly into the cylinder xDI obtained at 2000 RPM and throttle opening of 20% are shown in

**Figure 13.** The traces of torque and specific fuel consumption in a function of the fraction of fuel injected directly into

**•** For the throttle opening equal to 20% and engine speed 2000 RPM best results of specific fuel consumption and torque was observed for the ratio of fuel injected directly into the cylinder amounting to 0.62. In the described case, the mentioned operating parameters of the engine received a significant improvement in relation to the situation when the whole

Figure 14 shows charts of engine total efficiency and relative increase of the engine total efficiency ΔηDI + MPI for dual-injection operation in relation to operation with indirect fuel injection developed on the basis of the results of Figure 12 and Figure 13. The traces shown on Figure 14 are the result of parabolic approximation the points obtained by the calculations.

The engine total efficiency is determined by the formula (1). For the calculation the calorific

<sup>6</sup> 3.6 10

<sup>×</sup> <sup>=</sup> <sup>×</sup> (1)

*<sup>d</sup> BSFC W*

The highest increase of the total efficiency ΔηDI+MPI shown in the Figure 14 amounted to 4.58% for the first case and 2.18% in the second test point. In the first case the best efficiency was

*tot*

h

the cylinder xDI obtained for the throttle opening 20% and engine rotational speed 2000 RPM

amount of the fuel is injected into intake channels.

value of petrol Wd = 44 000 kJ / kg was assumed [13].

cylinder.

64 Advances in Internal Combustion Engines and Fuel Technologies

Figure 13.

**•** During the described above tests of the engine using gas analyzer Arcon Oliver K-4500 volumetric concentrations of individual exhaust components in the exhaust manifold were measured The concentration of carbon monoxide CO, carbon dioxide CO2, nitric oxide NO, unburned hydrocarbons HC and additionally exhaust gas temperature texh were investigat‐ ed. The total concentration of hydrocarbons in the exhaust HC was converted by the gas analyzer to hexane.

In Figure 15 registered at speed 2000 RPM and at throttle opening of 13% the traces of volumetric concentrations of the above mentioned chemicals and the exhaust gas temperature were shown depending on the fraction of fuel injected directly into the cylinder.

**•** The analysis of Figure 15 shows that with increase in the fraction of fuel injected directly into the cylinder the concentrations of carbon monoxide and hydrocarbons increase slightly, while the concentrations of nitrogen oxide and carbon dioxide decrease. Also the tempera‐ ture of gas leaving the engine cylinders decreased slightly. The difference between the NO concentration for injection only into the intake channel and only with direct injection into the cylinder is not high and is approximately 170 ppm. The concentration of HC for direct

was 14° CA before TDC. The measured absolute pressure in the intake manifold equalled 0.079 MPa. The direct injection pressure was set at 8 MPa, and the angle of start of injection was 281° CA before TDC. The fraction of fuel injected directly into the cylinder in dual-injection mode was equal to 0.62. For such a value the minimum of specific fuel consumption for those

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The tests were conducted to determine the differences in the combustion process in the engine for indirect fuel injection and for dual-injection with predetermined fraction of fuel injected directly into the cylinder providing minimum specific fuel consumption. An optoelectronic pressure sensor Optrand C82255-SP attached to a specially prepared spark plug and an angular incremental encoder Omron E6B-CWZ3E were used for this purpose. The data from both of the sensors were recorded using a portable PC with National Instruments DAQCard-6062 card

The indicator diagrams obtained for the operation with only indirect injection and using the

**Figure 17.** The comparison of the closed indicator diagrams for indirect injection and for the dual-injection with 62%

The increased surface area of the graph representing the positive work of the engine cycle is visible. The peak combustion pressure reached a value of 4.23 MPa at 21° CA after TDC with indirect injection and 4.60 MPa at 19.5° CA after TDC in the dual-injection mode. The peak combustion pressure with dual-injection is thus higher by the value of 0.37 MPa as compared with the result obtained for the injection only to the intake channels. In order to more precisely determine the differences resulting from the course of the indicator diagrams the indicated mean effective pressure IMEP was calculated based on the recorded data, respectively for the

of fuel injected directly into the cylinder, engine speed 2000 RPM, throttle opening of 20%

working with the application created in LabView environment.

dual-injection system were illustrated in Figure 17.

conditions was recorded.

**Figure 15.** The temperature and the volumetric concentrations of selected exhaust gas components obtained at 2000 RPM with the throttle opening 13%

**Figure 16.** The diagrams of temperature and concentrations of selected exhaust components obtained at the engine speed equal to 2000 RPM and 20% throttle opening

injection in a similar comparison is increased somewhat more, but without reaching the particularly high value - approximately 290 ppm.

**•** The following Figure 16 shows recorded at a speed of 2000 RPM and a throttle opening of 20% traces of the temperature and the concentrations of the previously mentioned exhaust gas components.

The character of changes in the parameters presented in Figure 16 is not significantly different from those observed in the previous case.

#### **3.3. Impact of the use of the dual-injection system on the combustion process**

In the second part of the experimental studies for engine speed 2000 RPM, throttle opening 20% and stoichiometric composition of the mixture the waveforms of an indicated pressure have been recorded. As in previously carried out studies in these conditions, ignition timing was 14° CA before TDC. The measured absolute pressure in the intake manifold equalled 0.079 MPa. The direct injection pressure was set at 8 MPa, and the angle of start of injection was 281° CA before TDC. The fraction of fuel injected directly into the cylinder in dual-injection mode was equal to 0.62. For such a value the minimum of specific fuel consumption for those conditions was recorded.

The tests were conducted to determine the differences in the combustion process in the engine for indirect fuel injection and for dual-injection with predetermined fraction of fuel injected directly into the cylinder providing minimum specific fuel consumption. An optoelectronic pressure sensor Optrand C82255-SP attached to a specially prepared spark plug and an angular incremental encoder Omron E6B-CWZ3E were used for this purpose. The data from both of the sensors were recorded using a portable PC with National Instruments DAQCard-6062 card working with the application created in LabView environment.

The indicator diagrams obtained for the operation with only indirect injection and using the dual-injection system were illustrated in Figure 17.

**Figure 17.** The comparison of the closed indicator diagrams for indirect injection and for the dual-injection with 62% of fuel injected directly into the cylinder, engine speed 2000 RPM, throttle opening of 20%

injection in a similar comparison is increased somewhat more, but without reaching the

**Figure 16.** The diagrams of temperature and concentrations of selected exhaust components obtained at the engine

**Figure 15.** The temperature and the volumetric concentrations of selected exhaust gas components obtained at 2000

**•** The following Figure 16 shows recorded at a speed of 2000 RPM and a throttle opening of 20% traces of the temperature and the concentrations of the previously mentioned exhaust

The character of changes in the parameters presented in Figure 16 is not significantly different

In the second part of the experimental studies for engine speed 2000 RPM, throttle opening 20% and stoichiometric composition of the mixture the waveforms of an indicated pressure have been recorded. As in previously carried out studies in these conditions, ignition timing

**3.3. Impact of the use of the dual-injection system on the combustion process**

particularly high value - approximately 290 ppm.

from those observed in the previous case.

speed equal to 2000 RPM and 20% throttle opening

gas components.

RPM with the throttle opening 13%

66 Advances in Internal Combustion Engines and Fuel Technologies

The increased surface area of the graph representing the positive work of the engine cycle is visible. The peak combustion pressure reached a value of 4.23 MPa at 21° CA after TDC with indirect injection and 4.60 MPa at 19.5° CA after TDC in the dual-injection mode. The peak combustion pressure with dual-injection is thus higher by the value of 0.37 MPa as compared with the result obtained for the injection only to the intake channels. In order to more precisely determine the differences resulting from the course of the indicator diagrams the indicated mean effective pressure IMEP was calculated based on the recorded data, respectively for the two cases. The method of numerical integration of relevant areas of the graphs of Figure 17 was applied. In order to provide increased accuracy trapezoid method was used.

The brake mean effective pressure BMEP was determined according to the formula (2) for both considered fuel systems :

$$BMEP\ = \frac{\pi \cdot \pi \cdot T}{500 \cdot V\_{ss}}\tag{2}$$

However, based on equation (3) it was possible to calculate the thermal efficiency of the engine in both cases:

$$\eta\_{thr} = \frac{N\_i}{N\_c} = \frac{30 \cdot MEPP \cdot V\_{ss} \cdot n}{\text{G}\_e \cdot \text{W}\_d} \tag{3}$$

**Figure 18.** The rate of pressure rise as a function of crank angle obtained for both of the considered fuel systems

occurring of peak rates of pressure rise typically higher than 0.5 MPa /° CA [14].

others in [15].

the combustion process.

The second stage of the analysis of the cylinder pressure charts obtained for both of the fuel systems was focused on the identifying of the process of mixture combustion. The method of the analysis of indicator diagram allowing to determine the mass fraction burned (MFB) in the cylinder as a function of crank angle was applied. This method is widely described among

Figure 19 shows the traces of the mass fraction burned as a function of crank angle obtained for both fuel systems. In the Figure 26, the ordinate grid lines corresponding to mass fraction burned in the cylinder of 0.1 and 0.9 are in bold. The mentioned values are important due to

The value of the angle of flame propagation is determined by the moment in which mass

 a

The fast burn angle Δαs is defined with the formula (5), as a difference between the angle of

(4)

*<sup>r</sup>* 10% *ign* D= aa

90% mass fraction burned - α90% and the angle of 10% mass fraction burned - α10%.

fraction burned equals to 10%, according to the formula (4):

**•** The analysis of the results indicates an increase of rate of pressure rise in the case of dualinjection of fuel. The peak rate of pressure rise amounted to 0.181 MPa/° CA for fuel injection into the intake channels and 0.253 MPa/° CA for dual-injection of fuel. The increase of the rate of pressure rise is not a favourable phenomenon, as it provides increased load in the cranktrain, however, the value obtained for the dual-injection system is not high. It is worth to mention that the occurrence of knock in a spark-ignition engine is characterized by

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The results of calculations of the brake mean effective pressure, the engine thermal efficiency and the indicated mean effective pressure were presented in the Table 2.


**Table 2.** Comparison of the indicators of work of the engine obtained with multipoint fuel injection and with dualinjection of fuel

Using dual-injection system about 2.6% increase in the indicated mean effective pressure and about 3.8% increase in the thermal efficiency were achieved compared to injection only into intake channels. These values are similar to those obtained in the corresponding comparison made for the specific fuel consumption for the considered engine operating conditions. On this basis, it can be concluded that the increase in indicated mean effective pressure and thermal efficiency shows improved combustion efficiency of the mixture prepared by dual-injection system. This fact can be explained as reflected in the simulations intensifying turbulence of the charge when part of the fuel is injected directly into the cylinder.

The last indicator in this part of the analysis of the indicator diagrams is the rate of pressure rise dpc/dα. The curve of this parameter as a function of crank angle was shown in Figure 18 for the crucial part of the indicator diagram. The rate of pressure rise was adopted as the primary indicator of the possibility of occurring of the knock combustion.

two cases. The method of numerical integration of relevant areas of the graphs of Figure 17

The brake mean effective pressure BMEP was determined according to the formula (2) for both

500 *ss*

However, based on equation (3) it was possible to calculate the thermal efficiency of the engine

30 *<sup>i</sup> ss*

*N IMEP V n N GW*

The results of calculations of the brake mean effective pressure, the engine thermal efficiency

*c ed*

*thr*

and the indicated mean effective pressure were presented in the Table 2.

BMEP [MPa] 0.745 0.769 3.22

IMEP [MPa] 0.931 0.955 2.585

Thermal efficiency ηthr [-] 0.395 0.410 3.797

charge when part of the fuel is injected directly into the cylinder.

primary indicator of the possibility of occurring of the knock combustion.

**Table 2.** Comparison of the indicators of work of the engine obtained with multipoint fuel injection and with dual-

Using dual-injection system about 2.6% increase in the indicated mean effective pressure and about 3.8% increase in the thermal efficiency were achieved compared to injection only into intake channels. These values are similar to those obtained in the corresponding comparison made for the specific fuel consumption for the considered engine operating conditions. On this basis, it can be concluded that the increase in indicated mean effective pressure and thermal efficiency shows improved combustion efficiency of the mixture prepared by dual-injection system. This fact can be explained as reflected in the simulations intensifying turbulence of the

The last indicator in this part of the analysis of the indicator diagrams is the rate of pressure rise dpc/dα. The curve of this parameter as a function of crank angle was shown in Figure 18 for the crucial part of the indicator diagram. The rate of pressure rise was adopted as the

h

*V*

× × <sup>=</sup> <sup>×</sup> (2)

× ×× = = <sup>×</sup> (3)

**xDI = 0 (MPI) xDI = 0.62 (MPI+DI) Increase from xDI=0, [%]**

was applied. In order to provide increased accuracy trapezoid method was used.

*<sup>T</sup> BMEP*

p t

considered fuel systems :

68 Advances in Internal Combustion Engines and Fuel Technologies

in both cases:

injection of fuel

**Figure 18.** The rate of pressure rise as a function of crank angle obtained for both of the considered fuel systems

**•** The analysis of the results indicates an increase of rate of pressure rise in the case of dualinjection of fuel. The peak rate of pressure rise amounted to 0.181 MPa/° CA for fuel injection into the intake channels and 0.253 MPa/° CA for dual-injection of fuel. The increase of the rate of pressure rise is not a favourable phenomenon, as it provides increased load in the cranktrain, however, the value obtained for the dual-injection system is not high. It is worth to mention that the occurrence of knock in a spark-ignition engine is characterized by occurring of peak rates of pressure rise typically higher than 0.5 MPa /° CA [14].

The second stage of the analysis of the cylinder pressure charts obtained for both of the fuel systems was focused on the identifying of the process of mixture combustion. The method of the analysis of indicator diagram allowing to determine the mass fraction burned (MFB) in the cylinder as a function of crank angle was applied. This method is widely described among others in [15].

Figure 19 shows the traces of the mass fraction burned as a function of crank angle obtained for both fuel systems. In the Figure 26, the ordinate grid lines corresponding to mass fraction burned in the cylinder of 0.1 and 0.9 are in bold. The mentioned values are important due to the combustion process.

The value of the angle of flame propagation is determined by the moment in which mass fraction burned equals to 10%, according to the formula (4):

$$
\Delta \alpha\_r = \alpha\_{10\%} - \alpha\_{ign} \tag{4}
$$

The fast burn angle Δαs is defined with the formula (5), as a difference between the angle of 90% mass fraction burned - α90% and the angle of 10% mass fraction burned - α10%.

**Figure 19.** The Mass Fraction Burned of the cylinder charge as a function of crank angle for MPI – fuel supply and for dual-injection of fuel (description in the text)

$$
\Delta \alpha\_s = \alpha\_{90\%} - \alpha\_{10\%} \tag{5}
$$

of complete combustion Δαo, which is the sum of the two above mentioned, has reached values, respectively 38.3° CA at indirect fuel injection and 35.4° CA for the dual-injection of fuel. This gives a reduction in the angle at which the most important part of the combustion process takes place of 2.9° CA i.e. about 7.6%. This is undoubtedly the reason for an increase in indicated mean effective pressure IMEP and thermal efficiency ηthr, which were analyzed above. The combustion of the mixture in a shorter time results lower heat losses occurring by the cylinder sleeve, because in this case a part of the cylinder sleeve in contact with a hot charge

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On the Figure 20 curves of speed of the charge combustion dMFB/dα as a function of crank angle were shown for the two fuel systems. The speed of the charge combustion was obtained by differentiating the mass fraction burned MFB shown in Figure 19 relative to crank angle.

**Figure 20.** The speed of combustion of the charge dMFB/dα in a function of the crank angle for both of the injection

The speed of the charge combustion in the most part of the period of fast burn achieved higher values of average 0.54% mass of the burned charge per 1° CA for dual-injection of fuel. The absolutedifferenceinthespeedofchargecombustionobtainedwithdual-injectionoffuelreaches a maximum value of 1.76% of the mass per 1° CA at 373.5° CA. In the second part of the period a fast burn with indirect fuel injection the process runs more intense, but the greatest effect on improving the thermal efficiency of the engine has increasing of the speed of the charge combustion in the first stage of the process, i.e. to reaching of 50% mass fraction burned [16].

Therefore, the above considerations represent a confirmation of the positive impact of using of the dual-injection system on the combustion process for the assumed engine operating conditions. The result of this interaction is improvement of the engine operation indicators, such as, among others Indicated mean effective pressure IMEP and thermal efficiency ηthr,

which values have a direct impact on the total efficiency of the engine ηtot.

has a smaller surface area.

systems

**•** The angle of complete combustion Δα<sup>o</sup> was defined as sum of the flame propagation angle Δαr and the fast burn angle Δαs - formula (6).

$$
\Delta \alpha\_o = \Delta \alpha\_r + \Delta \alpha\_s \tag{6}
$$

**•** The values of the angles characterizing the combustion process, which were indicated in Figure 26, were given in Table 3 respectively for indirect fuel injection and for dual-injection with 62% fraction of fuel injected directly into the cylinder.


**Table 3.** The values of the angles characterizing the combustion process

In the case of the dual-injection the angle of flame propagation was reduced from 17 to 16.5° CA, and, more importantly, the fast burn angle was decreased from 21.3 to 18.9° CA. The angle of complete combustion Δαo, which is the sum of the two above mentioned, has reached values, respectively 38.3° CA at indirect fuel injection and 35.4° CA for the dual-injection of fuel. This gives a reduction in the angle at which the most important part of the combustion process takes place of 2.9° CA i.e. about 7.6%. This is undoubtedly the reason for an increase in indicated mean effective pressure IMEP and thermal efficiency ηthr, which were analyzed above. The combustion of the mixture in a shorter time results lower heat losses occurring by the cylinder sleeve, because in this case a part of the cylinder sleeve in contact with a hot charge has a smaller surface area.

On the Figure 20 curves of speed of the charge combustion dMFB/dα as a function of crank angle were shown for the two fuel systems. The speed of the charge combustion was obtained by differentiating the mass fraction burned MFB shown in Figure 19 relative to crank angle.

*<sup>s</sup>* 90% 10% D= aa

*ors* D =D +D aaa

Δαr and the fast burn angle Δαs - formula (6).

70 Advances in Internal Combustion Engines and Fuel Technologies

dual-injection of fuel (description in the text)

with 62% fraction of fuel injected directly into the cylinder.

**Table 3.** The values of the angles characterizing the combustion process

 a

**Figure 19.** The Mass Fraction Burned of the cylinder charge as a function of crank angle for MPI – fuel supply and for

**•** The angle of complete combustion Δα<sup>o</sup> was defined as sum of the flame propagation angle

**•** The values of the angles characterizing the combustion process, which were indicated in Figure 26, were given in Table 3 respectively for indirect fuel injection and for dual-injection

**No Angle of Symbol MPI [°CA] 0.62DI [°CA] Difference to MPI [°CA]**

In the case of the dual-injection the angle of flame propagation was reduced from 17 to 16.5° CA, and, more importantly, the fast burn angle was decreased from 21.3 to 18.9° CA. The angle

 Ignition αign 346 346 0 10% mass fraction burned α10% 363 362.5 -0.5 90% mass fraction burned α90% 384.3 381.4 -2.9 Flame propagation Δα<sup>r</sup> 17 16.5 -0.5 Fast burn Δα<sup>s</sup> 21.3 18.9 -2.4 Complete combustion Δα<sup>o</sup> 38.3 35.4 -2.9

(5)

(6)

**Figure 20.** The speed of combustion of the charge dMFB/dα in a function of the crank angle for both of the injection systems

The speed of the charge combustion in the most part of the period of fast burn achieved higher values of average 0.54% mass of the burned charge per 1° CA for dual-injection of fuel. The absolutedifferenceinthespeedofchargecombustionobtainedwithdual-injectionoffuelreaches a maximum value of 1.76% of the mass per 1° CA at 373.5° CA. In the second part of the period a fast burn with indirect fuel injection the process runs more intense, but the greatest effect on improving the thermal efficiency of the engine has increasing of the speed of the charge combustion in the first stage of the process, i.e. to reaching of 50% mass fraction burned [16].

Therefore, the above considerations represent a confirmation of the positive impact of using of the dual-injection system on the combustion process for the assumed engine operating conditions. The result of this interaction is improvement of the engine operation indicators, such as, among others Indicated mean effective pressure IMEP and thermal efficiency ηthr, which values have a direct impact on the total efficiency of the engine ηtot.
