JP4839335B2 - In-cylinder injection engine - Google Patents

Description

  The present invention relates to an in-cylinder injection engine provided with a fuel injection valve that directly injects fuel into a combustion chamber.

  As a conventional in-cylinder injection engine, it has a variable valve timing mechanism that can change the opening and closing periods of the intake and exhaust valves, and the new air blow-off amount is estimated from the oxygen concentration in the exhaust passage detected by the air-fuel ratio sensor. Some control the valve overlap amount in accordance with the amount of blow-through, thereby reducing the in-cylinder residual gas amount and improving the filling efficiency (for example, see Patent Document 1).

  The gas combusted in the cylinder is pushed out by the piston in the exhaust stroke, but the volume of the clearance of the cylinder head cannot be scavenged, and high-temperature gas remains in the combustion chamber. When high-temperature gas remains in the combustion chamber, in a non-supercharged engine not equipped with a supercharger, fresh air is warmed by the residual gas, the air density is lowered, and the charging efficiency is lowered. In a supercharged engine equipped with a supercharger, in order to avoid knocking in an operation region of a medium load or higher, it is necessary to ignite after the optimal ignition timing of output and fuel consumption, and output and fuel consumption deteriorate.

  Therefore, in order to reduce the influence of the residual gas, there is a method of pushing out the residual gas from the combustion chamber by blowing new air from the intake port to the exhaust port as in Patent Document 1. However, this method requires a variable mechanism for the intake and exhaust valves, and the effect of pushing out the residual gas is weak under conditions where the pressure difference between the intake pipe and the exhaust pipe is small.

  As another method, there is a method of injecting fuel directly into the combustion chamber and utilizing the evaporative cooling effect of the fuel. This method uses sensible heat and latent heat generated when the fuel is heated and vaporized by injecting the fuel toward the residual gas. There is no need for a variable mechanism for the intake and exhaust valves, and the application range is not limited. There are features.

JP 2007-263083 A

  In the method using the evaporative cooling effect of fuel, when the intake valve opens and fresh air flows into the combustion chamber, the residual gas is diluted with fresh air. It is necessary to inject early. However, if all the fuel is injected at the beginning of the intake stroke at a time, the distance from the fuel injection valve to the piston is short, so that fuel adhesion to the piston increases. Since the fuel adhering to the piston is vaporized not by the heat of the residual gas but by the heat of the piston, the residual gas cannot be cooled.

  An object of the present invention is to solve the above-described problems, and in-cylinder injection engine that efficiently cools the residual gas in the fuel chamber without causing knocking in an operation region of a medium load or higher and improves output and fuel consumption. Is to provide.

  In order to achieve the above object, an in-cylinder injection engine according to the present invention includes a fuel injection valve disposed at a position where fuel can be directly injected into a combustion chamber, and a cylinder provided with control means for controlling opening and closing of the fuel injection valve. In the internal injection engine, the control means sets a fuel injection control range in which fuel is divided and injected two or more times during an intake stroke in one cycle in an operation region of medium load or more, and the injection division The number of times is controlled so as to increase in an operation region where the engine speed and the charging efficiency of the combustion chamber are small, and to decrease in an operation region where at least one of the engine speed and the charging efficiency is high.

  According to the present invention, it is possible to efficiently cool the residual gas in the combustion chamber and improve the output and fuel consumption without causing knocking in an operation region of medium load or higher.

  The present invention divides and sprays the fuel sprayed during the intake stroke of one cycle, and reduces the fuel injection amount sprayed at the initial stage of the intake stroke with respect to the total fuel injection amount sprayed during one intake stroke. Thus, the sprayed fuel is prevented from adhering to the piston, and the fuel is efficiently sprayed on the residual gas. That is, in the present invention, in order to reduce the fuel injection amount at the initial stage of the intake stroke, the number of divided fuels injected during the intake stroke is increased based on the operating conditions, thereby reducing the amount of sprayed fuel at the initial stage of the intake stroke. Control or control is performed to reduce the amount of sprayed fuel in the initial stage of the intake stroke and to increase the amount of sprayed fuel in the middle to later stages of the intake stroke as compared with the initial sprayed fuel. In the following embodiments, the residual gas is more efficiently obtained by performing both the determination of the number of divisions according to the operating conditions and the increase in the amount of fuel sprayed for the second and subsequent times from the amount of fuel sprayed for the first time. Next, a mode of spraying fuel will be described.

  FIG. 1 is a cross-sectional view showing a direct injection engine according to the first embodiment. The in-cylinder injection engine of this embodiment includes a supercharger. As shown in FIG. 1, the in-cylinder injection engine includes a cylinder head 1, a cylinder block 2, and a piston 3 inserted into the cylinder block 2, and a combustion chamber 30 is formed. 4 is provided. An intake port 5 and an exhaust port 6 are opened in the combustion chamber 30, and an intake valve 7 and an exhaust valve 8 that open and close each opening are provided. The intake valve 7 and the exhaust valve 8 are commonly used cam operation type valves. The intake valve 7 is closed at the top dead center and the exhaust valve 8 is opened at the top dead center. The opening period of the intake valve 7 is 220 deg (the crank angle when the crankshaft 11 is at the top dead center is 0 deg).

  A fuel injection valve 9 for directly injecting fuel into the combustion chamber 30 is provided on the intake port 5 side of the combustion chamber 30. As shown in FIG. 2, the fuel injection valve 9 is a multi-hole injector that is injected from each of six injection holes. FIG. 3 shows a cross-sectional shape 30 mm below (A-A line) from the tip of the nozzle hole of the sprays 24 a to 24 f injected from the nozzle hole. The sprays 24a to 24f are shaped so as not to come into contact with the intake valve 7 even when fuel is injected under the condition that the lift amount of the intake valve 7 is maximized (the intake valve is fully opened). The fuel injection valve 9 is arranged so that The fuel is pressurized by a high-pressure fuel pump (not shown) and injected into the combustion chamber 30 through the fuel injection valve 9. The fuel injection pressure is calculated from a fuel pressure sensor provided in a fuel pipe between the fuel injection valve 9 and the high-pressure fuel pump.

  The piston 3 is connected to the crankshaft 11 via a connecting rod 10. A crank angle sensor 12 for calculating the engine speed is attached to the crankshaft 11. The cylinder block 2 is provided with a water temperature sensor 13 for calculating the temperature of the engine cooling water and a knock sensor 14 for calculating the vibration frequency of the engine.

  A collector 15 is connected upstream of the intake pipe 5 </ b> A connected to the intake port 5, and the amount of fresh air to be sucked is adjusted by a throttle valve 16. The throttle valve 16 is, for example, an electronic control type, and operates according to the depression amount of an accelerator (not shown), and is set so that the opening degree increases as the depression amount increases. Although only one cylinder is shown in FIG. 2, in the present embodiment, air is distributed from the collector 15 to each cylinder in a four-cylinder engine having a displacement of 500 cc per cylinder and a compression ratio of 11.

  An intercooler 17 and a compressor 18 are provided upstream of the collector 15. The compressor 18 is connected to the turbine 20, and can rotate fresh air at atmospheric pressure or higher into the engine (collector 15) by rotating the turbine 20 with the energy of the exhaust gas. Since the supercharged air has a higher temperature and a lower density, it is cooled by the intercooler 17. An intake pressure sensor 19 is attached to the collector 15, and the intake pipe pressure (supercharging pressure) can be calculated.

  An exhaust temperature sensor 21 is attached downstream of the exhaust pipe 6 </ b> A connected to the exhaust port 6, and a bypass flow path 22 is connected upstream of the turbine 20. The bypass passage 22 is provided with a control valve 23 that operates so as to open when the exhaust gas temperature exceeds a predetermined value. The exhaust gas is released to the bypass passage 22 to control the supercharging amount. A three-way catalyst, an air-fuel ratio sensor, and the like (not shown) are provided downstream of the turbine 20.

  An engine control unit (not shown), which is a control means, receives signals from sensors (crank angle sensor 12, water temperature sensor 13, knock sensor 14, intake pressure sensor 19, exhaust temperature sensor 21, fuel pressure sensor, etc.) It is connected to control each device (intake valve 7, exhaust valve 8, fuel injection valve 9, control valve 23, etc.) based on the signal, and the ROM in the ECU stores the engine speed, water temperature, exhaust temperature, empty The set values of various devices corresponding to the fuel ratio and collector pressure are recorded as map data.

  FIG. 4 shows fuel injection map data for the supercharging condition used in the in-cylinder injection engine of this embodiment. The horizontal axis represents the engine speed detected by the crank angle sensor 12, and the vertical axis represents the charging efficiency detected by the intake pressure sensor 19. However, the charging efficiency is expressed as supercharging pressure because the in-cylinder injection engine of this embodiment is equipped with a supercharger. N in the figure indicates the number of fuel injections (number of injection divisions) in one cycle, and is set so that the number of injections decreases as the engine speed and the supercharging pressure increase. For example, under the condition of an engine speed of 2000 r / min and a supercharging pressure of 0.2 bar, the fuel necessary for combustion is divided into four parts and injected.

Next, in the in-cylinder injection engine equipped with the supercharger shown in FIG.
Example 1
The first embodiment is an example in which the vehicle is operated under a condition (medium load operation condition) in which acceleration is started by depressing the accelerator during traveling. However, the throttle is not fully open. First, the rotational speed is calculated from the signal from the crank angle sensor 12, and the intake pipe pressure is calculated from the signal from the intake pressure sensor 19. In this embodiment, the engine speed is 2000 r / min, the intake pipe pressure is 1.2 bar (corresponding to a supercharging pressure of 0.2 bar), and the fresh air amount is calculated from the map data based on the engine speed and the intake pipe pressure. . The air-fuel ratio is determined from the map data based on the engine speed and the intake pipe pressure, and the fuel injection amount is calculated from the fresh air amount and the air-fuel ratio. In this embodiment, the fresh air amount per cycle, 712 mg, the air-fuel ratio is 12, and the target fuel injection amount is 59.3 mg.

  Similarly, from the engine speed and the intake pipe pressure, the injection division number, the ratio of each injection amount to be divided and the injection timing corresponding to the divided injection are determined from the map data of FIG. FIG. 5 shows a relationship between the injection timing and the injection pulse width with respect to the crank angle, and FIG. 6 shows a graph of each injection pulse width. As shown in FIG. 5, the number of injection divisions is 4, and the first time is 16%, the second time is 20%, the third time is 32%, and the fourth time is 32%. It is set so that the injection ratio after the second time is increased with respect to the first time. Accordingly, the target injection amount for each injection is 9.49 mg for the first time, 11.86 mg for the second time, 18.98 mg for the third time, and 18.98 mg for the fourth time.

  When the fuel pressure is calculated from the signal from the fuel pressure sensor, an injection pulse width that is a target fuel injection amount is calculated by a calculation formula recorded in advance in the ECU. As shown in FIG. 6, the ejection pulse widths T1 to T4 are T1 = 0.4 ms, T2 = 0.5 ms, T3 = 0.8 ms, and T4 = 0.8 ms. The injection start timing for the injection is determined from the map data of the engine speed and the number of injections. In this embodiment, the first time is 20 degATDC, the second time is 60 degATDC, the third time is 100 degATDC, and the fourth time is 140 degATDC.

  FIG. 7 is a cross-sectional view showing the inside of the combustion chamber 30 immediately after the first fuel injection. As shown in FIG. 7, at the beginning of the intake stroke, the intake valve 7 starts to open, the exhaust valve 8 is closed, and the residual gas 25 burned in the previous cycle remains in the combustion chamber 30. This residual gas is a high temperature gas of about 700 ° C., for example. The residual gas 25 is pushed toward the exhaust valve 8 when the intake valve 7 opens and fresh air enters. Since the atomized fuel 24 is injected at 20 degATDC which is the initial stage of the intake stroke, it collides with the residual gas 25 and cools the residual gas 25. The first fuel injection amount is reduced to 16% of the total (total fuel amount injected within one intake stroke), and fuel is injected, so the kinetic energy of the spray is small and the penetration force is weak. Therefore, the atomized fuel 24 hardly cools against the piston 3 and can cool the residual gas efficiently. In the configuration in which the fuel injection valve 9 is disposed near the intake valve 7, the atomized fuel 24 injected from the fuel injection valve 9 and the fresh air flowing in from the intake valve 7 (arrow 5a in the figure) interfere with each other. However, in this embodiment, the engine speed is not so high and the supercharging pressure is low, so that the influence of fresh air on the spray is small. In the present embodiment, the first injection timing is 20 degATDC, but the optimal injection timing varies depending on the characteristics of fuel spray and the stroke of the engine.

  FIG. 8 is a cross-sectional view showing the inside of the combustion chamber 30 immediately after the second fuel injection in the intake stroke. As shown in FIG. 8, when the fresh air 5 a flows into the combustion chamber 30 due to the lowering of the piston 3, the residual gas 25 is mixed with the fresh air 5 a, and the high temperature portion of the residual gas 25 that is the spray target is moved by the piston 3. Move away from the spray point. Therefore, by increasing the fuel injection amount of the second time to 20% of the whole and 16% of the first time to increase the penetration force of the sprayed fuel 24, the fuel sprayed fuel 24 reaches the residual gas 25 that moves away, and efficiently. Residual gas can be cooled. If all of the remaining fuel is injected by the second spray, the penetrating force of the sprayed fuel 24 is too strong, and the amount of adhesion to the piston 3 increases and the cooling effect decreases.

  FIG. 9 is a cross-sectional view showing the inside of the combustion chamber 30 immediately after the third injection of fuel in the intake stroke. As shown in FIG. 9, the piston 3 is lowered from the second injection timing, and the distance between the spray point and the high temperature portion of the residual gas 25 is long. Therefore, the third fuel injection amount is set to 32%, which is higher than the second fuel injection amount 20%, and the penetration force of the spray fuel 24 is increased.

In the fourth spray fuel 24 in the intake stroke, the piston position is lowered from the third injection timing, and the distance between the spray point and the high temperature portion of the residual gas 25 is increased. However, when the injection timing is late, the time from the injection to the ignition timing is short, and increasing the fuel injection amount makes it difficult to form a uniform air-fuel mixture that is optimal for combustion. % Injection rate.
(Example 2)
The relationship between the injection timing and the injection pulse width with respect to the crank angle in the second embodiment is shown in FIGS. The operating conditions of this example are the same as those of Example 1. The difference from the first embodiment is that the injection ratio of each fuel injection amount in the intake stroke is 16% for the first time, 24% for the second time, 36% for the third time, and 34% for the fourth time. There are few points. The injection pulse width is T1 = 0.4 ms, T2 = 0.6 ms, T3 = 0.9 ms, and T4 = 0.6 ms. In this embodiment, since the fuel ratio of the second and third times is increased with respect to the first embodiment, the adhesion to the piston 3 is increased and the effect of cooling the fresh air is slightly reduced. Vaporization takes place in a short period of time, and a longer time from injection to ignition can be secured, so that the air-fuel mixture is made uniform and the combustion speed is increased. Thereby, knock resistance can be improved as compared with the first embodiment.
(Example 3)
The relationship between the injection timing and the injection pulse width with respect to the crank angle in the third embodiment is shown in FIGS. The operating conditions are conditions where the accelerator is depressed more than in the first embodiment (medium load to high load operating conditions), the engine speed is 2500 r / min, and the intake pipe pressure is 1.4 bar (corresponding to a supercharging pressure of 0.4 bar). In this embodiment, the fresh air amount per cycle, 830 mg, the air-fuel ratio is 11.5, and the target fuel injection amount is 72.2 mg.

  Similar to the first embodiment, the number of injection divisions, the ratio of each injection quantity to be divided and the injection timing corresponding to the divided injection are determined from the map data from the engine speed and the intake pipe pressure. As shown in FIG. 12, the number of injection divisions in this embodiment is 3, which is 16% for the first time, 32% for the second time, and 52% for the third time. The target injection amount for each injection is 11.55 mg for the first time, 23.1 mg for the second time, and 37.54 mg for the third time.

  The injection pulse width that is the target fuel injection amount is calculated by the calculation formula recorded in the ECU, and each injection pulse width is T1 = 0.5 ms, T2 = 1.0 ms, and T3 = 1.6 ms as shown in FIG. . The injection start timing for each injection is determined from the map data of the engine speed and the number of injections, and the first time is 20 degATDC, the second time is 70 degATDC, and the third time is 120 degATDC. Further, each injection interval (between T1 and T2, between T2 and T3, and between T3 and T4) is set to 2 ms or more because it requires a voltage boosting time to be supplied to the fuel injection valve 9.

In this embodiment, since the engine speed is higher than in the first and second embodiments, the flow velocity around the intake valve caused by the inflow of fresh air becomes faster, and the intake pipe pressure is also higher, so the fresh air density is higher. Is getting bigger. Therefore, if the injected fuel is divided into four as in the first and second embodiments, the momentum (penetrating force) of each sprayed fuel 24 becomes small, and the residual gas 25 cannot reach the fresh air. In the present embodiment, by reducing the number of divisions to 3, the influence of the fresh air flow around the intake valve can be reduced, and the residual gas 25 can be cooled by the sprayed fuel 24.
Example 4
14 and 15 show the relationship between the injection timing and the injection pulse width with respect to the crank angle in the fourth embodiment. The operating conditions of the present embodiment are high speed driving conditions (high load operating conditions) with the accelerator depressed, the engine speed is 3000 r / min, and the intake pipe pressure is 1.6 bar (corresponding to a supercharging pressure of 0.6 bar). In this embodiment, the amount of fresh air per cycle and cylinder is 949 mg, the air-fuel ratio is 11, and the target fuel injection amount is 86.3 mg.

  Similar to the first embodiment, the number of injection divisions, the ratio of each injection quantity to be divided and the injection timing corresponding to the divided injection are determined from the map data from the engine speed and the intake pipe pressure. As shown in FIG. 14, the number of injection divisions is 2, which is 16% for the first time and 84% for the second time. The target injection amount for each injection is 13.81 mg for the first time and 72.49 mg for the second time.

  The injection pulse width which becomes the target fuel injection amount is calculated by the calculation formula recorded in the ECU, and each injection pulse width is T1 = 0.59 ms and T2 = 3.11 ms as shown in FIG. The injection start timing for each injection is determined from the map data of the engine speed and the number of injections. The first injection start timing is 20 degATDC, and the second injection start timing is 60 degATDC.

  In this embodiment, since the engine speed and the intake pipe pressure are higher than those in the first to third embodiments, the number of divisions is made smaller than those in the first to third embodiments, that is, the number of divisions is set to two. The influence of the surrounding fresh air flow can be reduced, and the residual gas 25 can be cooled by the atomized fuel 24.

  As described above, according to the in-cylinder injection engine of the present embodiment, among the fuel injection amounts injected during one intake stroke, the fuel is injected divided into a plurality of times in order to reduce the fuel injected at the beginning of the intake stroke. As a result, the residual gas can be efficiently reduced in an operation region of a medium load or higher where knocking may occur. In particular, for each fuel spray injected in a divided manner, the first fuel is injected at the beginning of the intake stroke, and after the second time, a larger amount of sprayed fuel 24 is injected than the first, so that the spray is caused by the lowering of the piston 3. It is possible to reach the high-temperature residual gas 25 that moves away from the point, suppress the fuel adhesion to the piston 3, and cool the residual gas efficiently by the sensible heat and latent heat of the fuel. Further, the number of divisions can be reduced as the engine speed and charging efficiency increase, and the influence of fresh air flow on the fuel spray can be reduced. As a result, the knock resistance can be improved and the residual gas can be cooled. As a result, output and fuel consumption can be improved.

  A second embodiment will be described.

  FIG. 16 is a cross-sectional view showing a configuration of the in-cylinder injection engine of the present embodiment. As shown in FIG. 16, the in-cylinder engine of the present embodiment has no supercharger used in the in-cylinder injection engine of the previous embodiment. An air flow sensor 26 is provided upstream of the throttle valve 16. This embodiment differs from the engine configuration of FIG. 1 in that the charging efficiency is calculated based on the output signal of the airflow sensor 26. The fresh air amount is calculated from the signal from the air flow sensor 26 and the engine speed. Further, the charging efficiency is about 110% at the maximum, and the flow rate (fuel injection amount) required for the fuel injection valve 9 is designed to be smaller than that in the previous embodiment.

  FIG. 17 is fuel injection map data of the present embodiment. The horizontal axis indicates the engine speed detected by the crank angle sensor 12, and the vertical axis indicates the charging efficiency calculated by the airflow sensor 26. N in the figure indicates the number of fuel injections in one cycle, and is set so that the number of injections decreases as the engine speed and the charging efficiency increase.

(Example 5)
Next, in the in-cylinder injection engine of FIG. 16, the operation under one operating condition (engine speed and charging efficiency) will be described as a fifth embodiment.

  The fifth embodiment is a condition in which acceleration is started by depressing the accelerator during traveling, and the throttle is not fully opened. First, the engine speed is calculated from the signal of the crank angle sensor 12, and the fresh air amount is calculated from the signals of the air flow sensor 26 and the intake pressure sensor and the map data of the engine speed. In this embodiment, the engine speed is 2000 r / min, and the amount of fresh air per cylinder is 474 mg with a charging efficiency of 80%. The air-fuel ratio is determined from the map data of the engine speed and the charging efficiency, and the fuel injection amount is calculated from the fresh air amount and the air-fuel ratio. In this embodiment, the air-fuel ratio is 14, and the target fuel injection amount is 33.9 mg.

  From the engine speed and the charging efficiency, the number of injection divisions (map data in FIG. 17), the ratio of each injection quantity to be divided and the injection timing according to the divided injections are determined. The relationship between the injection timing and the injection pulse width with respect to the crank angle is shown in FIGS. As shown in FIG. 18, the number of injection divisions is 4, which is 16% for the first time, 20% for the second time, 32% for the third time, and 32% for the fourth time. Therefore, the target injection amount for each injection is 5.42 mg for the first time, 6.78 mg for the second time, 10.85 mg for the third time, and 10.85 mg for the fourth time.

  When the fuel pressure is calculated from the signal from the fuel pressure sensor, an injection pulse width that is a target fuel injection amount is calculated by a calculation formula recorded in advance in the ECU. As shown in FIG. 19, the ejection pulse widths are T1 = 0.64 ms, T2 = 0.81 ms, T3 = 1.32 ms, and T4 = 1.32 ms. The injection start timing for each injection is determined from the map data of the engine speed and the number of injections. In this embodiment, the first time is 20 degATDC, the second time is 60 degATDC, the third time is 100 degATDC, and the fourth time is 140 degATDC.

  This embodiment also has the same operational effects as the previous embodiment. That is, even in a cylinder injection engine without a supercharger, the residual gas can be efficiently cooled by dividing the spray and determining the injection pulse width according to the distance between the spray point and the residual gas. Further, in a cylinder injection engine without a supercharger, it is possible to improve the charging efficiency.

  As described above, the present invention is not limited to the above-described embodiment, and various other ones are assumed.

1 is a cross-sectional view showing a first embodiment of an in-cylinder injection engine according to the present invention. It is a side view which shows the fuel injection valve of FIG. 1, and a spray shape. It is sectional drawing which shows the spray shape in the AA of FIG. It is an injection frequency data map with respect to the engine speed and supercharging pressure of the first embodiment. 3 is a graph showing the relationship between fuel spray timing and intake valve lift amount in Example 1. It is a graph which shows the injection pulse width | variety of each injection time T1-T4 of FIG. 1 is a cross-sectional view showing a state in an engine immediately after a first fuel injection in Embodiment 1. FIG. FIG. 3 is a cross-sectional view showing a state in the engine immediately after the second fuel injection in the first embodiment. FIG. 3 is a cross-sectional view showing a state in the engine immediately after the third fuel injection in the first embodiment. It is a figure which shows the relationship between the fuel spray timing in Example 2, and an intake valve lift amount. It is a graph which shows the injection pulse width of each injection time T1-T4 of FIG. 6 is a graph showing the relationship between fuel spray timing and intake valve lift amount in Example 3. It is a graph which shows the injection pulse width of each injection time T1-T4 of FIG. It is a graph which shows the relationship between the fuel spray timing in Example 4, and an intake valve lift amount. It is a graph which shows the injection pulse width of each injection time T1-T4 of FIG. It is sectional drawing which shows 2nd Embodiment of the cylinder injection engine which concerns on this invention. It is an injection frequency data map with respect to the engine speed and filling efficiency of 2nd Embodiment. It is a graph which shows the relationship between the fuel spray timing in Example 5, and an intake valve lift amount. It is a graph which shows the injection pulse width of each fuel injection time T1-T4 of FIG.

Explanation of symbols

9 Fuel Injection Valve 12 Crank Angle Sensor 18 Compressor 19 Intake Pressure Sensor 24 Atomized Fuel 25 Residual Gas 30 Combustion Chamber

Claims (4)

A piston disposed in the combustion chamber, wherein an intake port for supplying air to the combustion chamber, a fuel injection valve arranged fuel injection possible positions directly into the combustion chamber, the supplied air and fuel into said combustion chamber In a cylinder injection engine comprising an ignition plug for igniting an air-fuel mixture and a control means for performing opening / closing control of the fuel injection valve,
The control means sets a fuel injection control range in which the fuel is divided into two or more injections during an intake stroke in one cycle in an operation region of medium load or more, and the number of injection divisions is increase charging efficiency of the engine speed and the combustion chamber in a small operation region, the combination with at least one of the engine speed and the charging efficiency is controlled to reduce the high operating region, the divided individual fuel injection amount is the An in-cylinder injection engine that controls the fuel injection valve so as to increase as the piston descends .   The in-cylinder injection engine according to claim 1, further comprising a supercharger that increases the charging efficiency of the combustion chamber.   The in-cylinder injection engine according to claim 1, wherein the first fuel injection is performed at an early stage of an intake stroke, and the second and subsequent fuel injections are performed until the intake valve is fully closed.   4. The in-cylinder injection engine according to claim 3, wherein the divided fuel injection amount is defined by a valve opening pulse width of the fuel injection valve.

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