JP2016098794A - Internal combustion engine control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To realize diesel combustion while suppressing the generation amount of smoke in an internal combustion engine using fuel the self-ignition temperature of which is relatively high.SOLUTION: According to the present invention, pre-ignition is implemented by spark ignition during a compression stroke. Thereafter, before a top dead center in the compression stroke, main injection is implemented by a cylinder injection valve. As a result, fuel injected by the main injection is started with a flame generated by the pre-ignition as a starting point, and self-ignition and diffusion combustion occur. According to the present invention, in an operation region where an engine load of an internal combustion engine is equal to or lower than a predetermined load, the fuel injection for the pre-ignition is implemented by pre-ignition by the cylinder injection valve. Further, in an operation region where the engine load of the internal combustion engine is higher than the predetermined load, the fuel injection for the pre-ignition is implemented by port injection by a port injection valve and the pre-ignition by the cylinder injection valve.SELECTED DRAWING: Figure 9

Description

  The present invention relates to a control device for an internal combustion engine.

  So-called diesel combustion, which is a combustion mode in which fuel is directly injected into compressed air in the combustion chamber and self-ignited and diffusely combusted, has higher thermal efficiency than combustion by spark ignition. In recent years, in order to enjoy such advantages of diesel combustion even in a gasoline engine, a technology for establishing combustion by self-ignition and diffusion combustion of gasoline has been developed.

  For example, in the technique disclosed in Patent Document 1, first, the first fuel injection is performed in the first half of the compression stroke by the in-cylinder injection valve, thereby forming a substantially homogeneous mixture throughout the combustion chamber. Then, spark ignition is performed on the air-fuel mixture formed by the first fuel injection. Thereafter, the injected fuel is burned by performing the second fuel injection, and the remaining fuel is self-ignited by the temperature and pressure rise in the combustion chamber due to this combustion.

  Patent Document 2 discloses a technique for realizing diesel combustion using natural gas or the like having a relatively high self-ignition temperature as a fuel. In the technique disclosed in Patent Document 2, first, an air-fuel mixture capable of spark ignition is formed by performing fuel injection in a predetermined spark ignition region in the combustion chamber in the initial stage or middle stage of the compression stroke. Then, spark ignition combustion is performed by igniting the air-fuel mixture formed in the spark ignition region at a time immediately before the top dead center of the compression stroke. As a result, the combustion chamber is brought into a high temperature and high pressure state capable of self-ignition of natural gas. Thereafter, the fuel is directly injected into the combustion chamber in a high-temperature and high-pressure state, and the fuel is diesel-combusted.

  Further, Patent Document 3 includes a port injection valve that injects fuel into an intake port and an in-cylinder injection valve that injects fuel into a cylinder. A mixture of fuel and air supplied into the cylinder is disclosed in Patent Document 3. In an internal combustion engine that burns by self-ignition, a technique for setting the fuel injection amount from the in-cylinder injection valve to 3 to 11% of the total fuel injection amount is disclosed.

JP 2002-276442 A JP 2003-254105 A JP 2011-236792 A

  When diesel combustion is performed using a fuel having a relatively high self-ignition temperature such as gasoline, if the amount of fuel injected into the combustion chamber increases as the engine load increases, the amount of oxygen relative to the fuel may be locally insufficient. As a result, the amount of smoke generated may increase. The present invention has been made in view of such a problem, and an object of the present invention is to realize diesel combustion while suppressing the amount of smoke generated in an internal combustion engine using a fuel having a relatively high self-ignition temperature. To do.

In the first invention, pre-combustion by spark ignition is performed during the compression process. Thereafter, main injection by the in-cylinder injection valve is executed before the top dead center of the compression stroke. Thereby, starting from the flame generated by the pre-combustion, combustion of the fuel injected by the main injection is started, and further, self-ignition and diffusion combustion of the fuel occur. In the present invention, fuel injection for pre-combustion is performed by pre-injection using the in-cylinder injection valve in an operation region where the engine load of the internal combustion engine is equal to or less than a predetermined load. In an operation region where the engine load of the internal combustion engine is higher than a predetermined load, fuel injection for pre-combustion is performed by port injection by the port injection valve and pre-injection by the in-cylinder injection valve.

  It should be noted that the terms “pre” and “main” used in the present invention are merely expressions representing the temporal context of each fuel injection or each combustion, other than those having the technical significance shown below. It should not be limited. In the following, the fuel injected by the pre-injection by the in-cylinder injection valve is referred to as “pre-injection fuel”, the fuel injected by the main injection by the in-cylinder injection valve is referred to as “main fuel injection”, and the port injection The fuel injected by the port injection by the valve may be referred to as “port injection fuel”.

  More specifically, the control device for an internal combustion engine according to the first aspect of the invention includes a cylinder injection valve capable of injecting fuel into a combustion chamber of the internal combustion engine, and a port injection valve capable of injecting fuel into an intake port of the internal combustion engine. An ignition device whose relative position with respect to the in-cylinder injection valve is determined so that the spray injected from the in-cylinder injection valve passes through the ignitable region and can be directly ignited, and an engine of the internal combustion engine In an operation region where the load is equal to or less than a predetermined load, pre-injection is executed by the in-cylinder injection valve during the compression stroke, and the pre-spray formed by the pre-injection is ignited by the ignition device. A first pre-combustion unit that performs a first pre-combustion by burning a part of the injected fuel, and an operation range in which the engine load of the internal combustion engine is higher than the predetermined load, during the exhaust stroke or the intake stroke, A gas mixture is formed by performing port injection with a fuel injection valve, and a pre-spray is formed in the gas mixture by performing pre-injection with the in-cylinder injection valve during the compression stroke. Second pre-combustion in which second pre-combustion is performed by burning part of the fuel injected by the first pre-injection and part of the fuel injected by the port injection by igniting the spray with the ignition device And a flame after the occurrence of the first pre-combustion or the second pre-combustion and before the compression stroke top dead center, starting from a flame generated by the first pre-combustion or the second pre-combustion By starting the execution of the main injection by the in-cylinder injection valve at the main injection timing set to start the combustion of the injected fuel, the fuel self-ignition is generated and at least The part of the fuel injected by the main injection is diffusion combustion and a main combustion section which performs main combustion.

  In the present invention, in an operating region where the engine load of the internal combustion engine is equal to or less than a predetermined load, first, pre-injection by the in-cylinder injection valve during the compression stroke and ignition by the ignition device to the pre-spray formed by the pre-injected fuel are performed. First pre-combustion is performed. Thereafter, the execution of the main injection is started before the top dead center of the compression stroke, whereby fuel self-ignition and diffusion combustion (main combustion) occur. The main injection is an injection that starts executing before the compression stroke top dead center. However, the execution may continue until the compression stroke top dead center.

  Here, the timings of the pre-injection, ignition, and main injection are set so that the combustion of the main injection fuel is started from the flame generated by the first pre-combustion. When the combustion of the main injected fuel is started, the temperature and pressure in the combustion chamber rise, so that fuel self-ignition occurs, and at least a part of the main injected fuel is diffusely burned. Further, some of the pre-injected fuel is combusted by ignition by the ignition device, and most of the fuel is combusted by self-ignition or diffusion combustion after the start of the main injection. Therefore, both the pre-injected fuel and the main injected fuel contribute to the output of the internal combustion engine. Therefore, it is possible to realize diesel combustion with high thermal efficiency.

Here, when the engine load of the internal combustion engine increases, it is necessary to increase the amount of fuel supplied into the combustion chamber. However, since the main injection is performed when the pressure in the combustion chamber near the top dead center of the compression stroke is high, the penetration of the fuel spray injected from the in-cylinder injection valve is reduced. That is, the fuel spray injected by the main injection is difficult to spread over a wide range. Therefore, if the amount of the main injection fuel is excessively increased, the oxygen present around the spray of the main injection fuel, that is, the amount of oxygen provided for the combustion of the main injection fuel becomes insufficient with respect to the fuel, As a result, the amount of smoke generated may increase.

  In addition, as described above, most of the pre-injected fuel remains unburned without being burned by ignition by the ignition device, and remains in the combustion chamber in an unburned state when main injection is executed. Therefore, if the amount of pre-injected fuel is increased instead of or together with the main injected fuel, the remaining amount of pre-injected fuel existing in the combustion chamber increases when the main injection is executed. However, if the pre-injected fuel is excessively increased, when the main injection is executed, the fuel concentration may become excessively high in a portion where the unburned residue of the pre-injected fuel and the main injected fuel overlap in the combustion chamber. is there. In this case, the amount of oxygen present around the portion, that is, the amount of oxygen used for combustion of the fuel present in the portion is insufficient with respect to the fuel, and as a result, the amount of smoke generated may increase. There is.

  That is, in both the pre-injection and the main injection, if the injection amount is excessively increased, the amount of smoke generated may increase. Therefore, in the present invention, in the operation region where the engine load of the internal combustion engine is higher than the predetermined load, instead of the first pre-combustion using only pre-injection, port injection by the port injection valve and pre-injection by the in-cylinder injection valve are performed. Second pre-combustion to be used together is performed. Here, the predetermined load is an engine load that requires a relatively large total fuel injection amount during one combustion cycle, and an excessive amount of smoke generated due to an increase in the pre-injection fuel amount or the main injection fuel amount. This is the threshold of the engine load that is expected to increase.

  When performing the second pre-combustion, port injection is performed by the port injection valve during the exhaust stroke or the intake stroke. That is, the port injection may be executed when the intake valve is closed, or may be executed when the intake valve is open. When the port injection is executed, an air-fuel mixture is formed in the combustion chamber by the port injection fuel. Next, pre-injection is executed by the in-cylinder injection valve during the compression stroke. As a result, a pre-spray is formed by the pre-injected fuel in the air-fuel mixture formed by the port injected fuel. The pre-spray is ignited by an ignition device. By this ignition, a part of the pre-injected fuel and a part of the port-injected fuel are combusted. Here, in the second pre-combustion, similarly to the first pre-combustion, it is sufficient to perform combustion to such an extent that a flame that serves as a fire type for main fuel combustion is formed. Therefore, in the present invention, the air-fuel ratio of the air-fuel mixture formed by the port-injected fuel is such that the port-injected fuel combusted by the propagation of flame generated by the ignition to the pre-spray by the ignition device is part of the port-injected fuel amount. The port injection fuel amount is set so that the air-fuel ratio can be suppressed.

  The port-injected fuel that remains unburned without being burned by the flame propagation caused by ignition by the ignition device is burned by the self-ignition after the main fuel injection or diffusion combustion together with the unburned residue of the pre-injection. Thereby, in addition to the main injection fuel and the pre-injection fuel, the port injection fuel also contributes to the output of the internal combustion engine. Therefore, even when the second pre-combustion is performed instead of the first pre-combustion, diesel combustion with high thermal efficiency can be realized.

Further, in the second pre-combustion according to the present invention, since the port injection is used in addition to the pre-injection, the pre-injected fuel can be reduced as compared with the first pre-combustion using only the pre-injection. Furthermore, the port injection fuel that forms the air-fuel mixture in the combustion chamber is more widely distributed in the combustion chamber than the main injection fuel. Therefore, at the time of main combustion, it is possible to sufficiently secure oxygen necessary for the combustion of the port injection fuel. Further, the port-injected fuel is distributed in a wider range in the combustion chamber than the unburned residue of the pre-injected fuel. Therefore, the fuel concentration in the portion where the main injection fuel and the port injection fuel overlap is lower than the fuel concentration in the portion where the main injection fuel and the pre-injection fuel (unburned fuel) overlap. Therefore, the port-injected fuel is unlikely to cause smoke as compared with either the main-injected fuel or the pre-injected fuel.

  For this reason, in the operation region where the engine load is higher than the predetermined load, that is, the operation region where the total fuel injection amount during one combustion cycle is relatively large, the second pre-combustion is performed instead of the first pre-combustion, so that the smoke Diesel combustion can be realized while suppressing the generation amount of NO.

  In the second invention, fuel injection for pre-combustion is performed only by port injection by the port injection valve.

  More specifically, the control device for an internal combustion engine according to the second invention includes a cylinder injection valve capable of injecting fuel into a combustion chamber of the internal combustion engine, and a port injection valve capable of injecting fuel into an intake port of the internal combustion engine. An ignition device whose relative position to the in-cylinder injection valve is determined so that the spray injected from the in-cylinder injection valve passes through the ignitable region and can be directly ignited, and an exhaust stroke or An air-fuel mixture is formed by executing port injection by the port injection valve during the intake stroke, and is injected by the port injection by igniting the air-fuel mixture by the ignition device in the combustion chamber during the compression stroke. A pre-combustion section that performs pre-combustion by burning a part of the fuel, and a timing after the occurrence of the pre-combustion and before the top dead center of the compression stroke, starting from the flame generated by the pre-combustion By starting execution of main injection by the in-cylinder injection valve at the main injection timing set so that combustion of the injected fuel is started, fuel self-ignition is generated and at least the fuel injected by the main injection A main combustion section for performing main combustion by diffusing and burning a part of the main combustion section.

  In the present invention, when pre-combustion is performed, an air-fuel mixture is formed in the combustion chamber by performing port injection by the port injection valve during the exhaust stroke or the intake stroke. At this time, the port injection may be executed when the intake valve is closed, or may be executed when the intake valve is open. And precombustion is performed by igniting with the ignition device with respect to the air-fuel | gaseous mixture formed in the combustion chamber with the port injection fuel. In this pre-combustion as well, it is only necessary to perform combustion to such an extent that a flame as a fire type for main fuel combustion is formed. Therefore, the air-fuel ratio of the air-fuel mixture formed by the port-injected fuel is such that a flame is generated when the fuel is ignited by ignition by the igniter, but in the propagation of the flame, only a part of the port-injected fuel burns. Thus, the port injection fuel amount is set.

  The port-injected fuel that remains unburned without being burned by the flame propagation caused by ignition by the ignition device is burned by the self-ignition after the main fuel injection or diffusion combustion together with the unburned residue of the pre-injection. Thereby, in addition to the main injected fuel, the port injected fuel also contributes to the output of the internal combustion engine. Therefore, even when pre-combustion according to the present invention is performed, diesel combustion with high thermal efficiency can be realized. Further, as described above, the port injection fuel is less likely to cause smoke than the main injection fuel as well as the pre-injection fuel.

  Therefore, according to the present invention, by performing pre-combustion using only port injection, diesel combustion can be realized while suppressing the amount of smoke generated.

In the present invention, the pre-combustion part may be a second pre-combustion part, and the pre-combustion may be a second pre-combustion. The control device for an internal combustion engine according to the present invention executes pre-injection by the in-cylinder injection valve during the compression stroke, and ignites the pre-spray formed by the pre-injection by an ignition device. You may further provide the 1st pre combustion part which burns a part of fuel injected by injection and performs 1st pre combustion. In this case, the first pre-combustion unit performs the first pre-combustion in the operation region where the engine load of the internal combustion engine is equal to or less than the predetermined load, and the second pre-combustion unit is included in the operation region where the engine load of the internal combustion engine is higher than the predetermined load. Second pre-combustion may be performed. The main combustion section starts after the first pre-combustion or the second pre-combustion and before the compression stroke top dead center, and the flame generated by the first pre-combustion or the second pre-combustion The main combustion may be performed by starting the execution of the main injection by the in-cylinder injection valve at the main injection timing set so that the combustion of the injected fuel is started.

  According to this, in the operation region where the engine load is higher than the predetermined load, that is, in the operation region where the total fuel injection amount in one combustion cycle is relatively large, the second pre-combustion is performed instead of the first pre-combustion. In addition, diesel combustion can be realized while suppressing the amount of smoke generated. Further, in an operation region where the engine load is lower than a predetermined load, that is, an operation region where the total fuel injection amount during one combustion cycle is relatively small, it becomes difficult to secure a sufficient amount of port injection fuel. There is a possibility that the air-fuel ratio of the air-fuel mixture formed by the fuel becomes excessively high. In this case, the ignitability of the fuel during pre-combustion is reduced, and as a result, misfire is likely to occur. According to the above, in the operation region where the engine load is lower than the predetermined load, the first pre-combustion using the pre-injection is performed by the first pre-combustion unit, thereby improving the ignitability of the fuel during the pre-combustion. it can.

  According to the present invention, in an internal combustion engine that uses fuel having a relatively high self-ignition temperature, diesel combustion can be realized while suppressing the amount of smoke generated.

It is a figure which shows schematic structure of the internal combustion engine with which the Example of this invention is applied, its intake system, and an exhaust system. FIG. 2 is a view for explaining the arrangement of spark plugs mounted on the internal combustion engine shown in FIG. 1. It is a figure for demonstrating the basic combustion control performed in the Example of this invention. It is a figure which shows transition of the heat release rate in a combustion chamber when the basic combustion control which concerns on the Example of this invention is performed. It is a figure which shows the correlation with the pre-injection fuel amount and the combustion efficiency of pre-injection fuel in the pre-injection in the basic combustion control which concerns on the Example of this invention. It is a figure which shows the change of transition of the heat generation ratio in a combustion chamber at the time of changing the ratio of the amount of pre-injection fuel and the main injection fuel quantity in the basic combustion control which concerns on the Example of this invention. It is a figure for demonstrating the high load combustion control performed in Example 1 of this invention. In the basic combustion control and the high load combustion control executed in the first embodiment of the present invention, it is a diagram showing the correlation between the engine load of the internal combustion engine and the thermal efficiency, and the correlation between the engine load of the internal combustion engine and the amount of smoke generated. is there. It is a flowchart which shows the control flow of the combustion control which concerns on Example 1 of this invention. It is a figure which shows the control map used for the combustion control which concerns on Example 1 of this invention. It is a figure for demonstrating the high load combustion control performed in Example 2 of this invention. It is a flowchart which shows the control flow of the combustion control which concerns on Example 2 of this invention. It is a figure which shows the control map used for the combustion control which concerns on Example 2 of this invention.

  Hereinafter, specific embodiments of the present invention will be described with reference to the drawings. The dimensions, materials, shapes, relative arrangements, and the like of the components described in the present embodiment are not intended to limit the technical scope of the invention to those unless otherwise specified.

  FIG. 1 is a diagram showing a schematic configuration of an internal combustion engine to which the present invention is applied and its intake / exhaust system. An internal combustion engine 1 shown in FIG. 1 is a 4-stroke cycle spark ignition internal combustion engine (gasoline engine) having a plurality of cylinders. In FIG. 1, only one cylinder among a plurality of cylinders is shown.

  Each cylinder 2 of the internal combustion engine 1 is internally slidably provided with a piston 3. The piston 3 is connected to an output shaft (crankshaft) (not shown) via a connecting rod 4. Further, the inside of the cylinder 2 communicates with the intake port 7 and the exhaust port 8. The opening end of the intake port 7 in the cylinder 2 is opened and closed by an intake valve 9. The open end of the exhaust port 8 in the cylinder 2 is opened and closed by an exhaust valve 10. The intake valve 9 and the exhaust valve 10 are each opened and closed by an intake cam and an exhaust cam (not shown).

  Each cylinder 2 is provided with an in-cylinder injection valve 61, a port injection valve 62, and a spark plug 5. The in-cylinder injection valve 61 is disposed at the central top of the combustion chamber formed in the cylinder 2 and directly injects fuel into the cylinder. The port injection valve 62 is provided in the intake port 7 and injects fuel in the direction of the opening end in the intake port 7. The spark plug 5 ignites the fuel (spray or mixture) in the combustion chamber of the cylinder 2.

  Here, the arrangement of the spark plug 5 in the cylinder 2 will be described with reference to FIG. The in-cylinder injection valve 61 according to the present embodiment has injection holes 61a so that fuel can be injected in 16 directions substantially radially as shown in FIG. Then, at least one of the fuel sprays injected from the nozzle hole 61a passes through the region 5a between the electrodes, which is an ignitable region of the spark plug 5, and in the region 5a with respect to the spray that has passed. The relative position of the spark plug 5 with respect to the in-cylinder injection valve 61, particularly the relative position of the region 5 a with respect to the in-cylinder injection valve 61, is determined so that direct ignition can be performed by the spark generated between the electrodes. The spark plug 5 and the in-cylinder injection valve 61 configured as described above can realize spray guide combustion. That is, the in-cylinder injection valve 61 and the spark plug 5 enable ignition of the fuel spray that passes through the region 5a at any time regardless of the opening timing of the intake valve 9 of the internal combustion engine 1 or the position of the piston 3. . In this embodiment, the spark plug 5 is located between the openings of the two intake ports 7 so as not to interfere with the operation of the intake valve 9 and the exhaust valve 10. However, the position of the ignition device according to the present invention is not limited to between the openings of the two intake ports.

  Returning to FIG. 1, the intake port 7 communicates with the intake passage 70. A throttle valve 71 is disposed in the intake passage 70. An air flow meter 72 is disposed in the intake passage 70 upstream of the throttle valve 71. On the other hand, the exhaust port 8 communicates with the exhaust passage 80. An exhaust purification catalyst 81 for purifying the exhaust discharged from the internal combustion engine 1 is disposed in the exhaust passage 80. As will be described later, the air-fuel ratio of the exhaust discharged from the internal combustion engine 1 is a lean air-fuel ratio that is higher than the stoichiometric air-fuel ratio. Therefore, as the exhaust purification catalyst 81, a selective reduction type NOx catalyst capable of purifying NOx in the lean air-fuel ratio exhaust or a filter capable of collecting particulate matter (PM) in the exhaust can be employed.

The internal combustion engine 1 is also provided with an electronic control unit (ECU) 20. The ECU 20 is a unit that controls the operating state of the internal combustion engine 1, an exhaust purification device, and the like. ECU20
The air flow meter 72, the crank position sensor 21, and the accelerator position sensor 22 are electrically connected to each other, and the detection values of the sensors are input to the ECU 20. Therefore, the ECU 20 is an internal combustion engine 1 such as an intake air amount detected by the air flow meter 72, an engine rotational speed calculated based on a detected value of the crank position sensor 21, and an engine load based on a detected value of the accelerator position sensor 22. It is possible to grasp the operating state of Further, the in-cylinder injection valve 61, the port injection valve 62, the spark plug 5, the throttle valve 71, and the like are electrically connected to the ECU 20, and each of these elements is controlled by the ECU 20.

<Basic combustion control>
The basic combustion control, which is the basic combustion control executed in the internal combustion engine 1 configured as described above, will be described with reference to FIG. FIG. 3 shows the flow of fuel injection and ignition related to basic combustion control performed in the internal combustion engine 1 (see the upper part of FIG. 3A), and the fuel injection and ignition in the time series from the left side to the right side of the figure. FIG. 4 schematically shows changes in events related to combustion that are assumed to occur in the combustion chamber (see the lower part of FIG. 3A). FIG. 3B shows a temporal correlation between pre-injection, main injection, and ignition, which are the fuel injections shown in FIG. Note that the form shown in FIG. 3 is shown only schematically for explaining the basic combustion control according to the present embodiment, and the present invention should not be interpreted as being limited to this form.

  In the basic combustion control according to the present embodiment, pre-injection and main injection are executed by the in-cylinder injection valve 61 in one combustion cycle. The pre-injection is a fuel injection that is performed during the compression stroke. The main injection is a fuel injection that is executed after the pre-injection and before the compression stroke top dead center (TDC). The execution of the main injection is started at a time before TDC, but the execution may be continued until after TDC. Then, as shown in FIG. 3B, the injection start timing of pre-injection (hereinafter simply referred to as “pre-injection timing”) is Tp, and the injection start timing of main injection (hereinafter simply referred to as “main injection timing”). Is called Tm. Further, an interval (Tm−Tp) between the pre-injection timing and the main injection timing is defined as a first injection interval Di1. Moreover, the combustion by pre-injection is performed as the spray guide combustion mentioned above. That is, ignition by the spark plug 5 is performed on pre-spray formed by fuel injected by pre-injection (hereinafter referred to as “pre-injected fuel”). This ignition timing is Ts as shown in FIG. 3B, and an interval (Ts−Tp) from the start of execution of pre-injection until ignition is defined is defined as an ignition interval Ds.

Next, the flow of basic combustion control according to the present invention will be described.
(1) Pre-injection In basic combustion control, pre-injection is first performed at pre-injection timing Tp during the compression stroke in one combustion cycle. Note that the pre-injection timing Tp is determined based on a correlation with a main injection timing Tm described later. By performing the pre-injection, as shown in FIG. 2, the pre-spray of the pre-injected fuel injected from the in-cylinder injection valve 61 passes through the ignitable region 5a of the spark plug 5 in the combustion chamber. Immediately after the execution of the pre-injection is started, the pre-spray of the pre-injected fuel does not diffuse widely in the combustion chamber, and the surrounding air is drawn in at the tip of the combustion chamber by the penetration force of the spray. Continue on. Therefore, a stratified mixture is formed in the combustion chamber by the pre-spraying of the pre-injected fuel.

(2) Ignition to pre-injected fuel Then, with respect to the pre-spray of the pre-injected fuel stratified as described above, the ignition plug 5 is operated at an ignition timing Ts when a predetermined ignition interval Ds has elapsed from the pre-injection timing. Ignition is performed. As described above, since the pre-injected fuel is stratified, the local air-fuel ratio around the spark plug 5 becomes an air-fuel ratio that can be combusted by the ignition even if the amount of the pre-injected fuel is small. Yes. By this ignition, spray guide combustion by the pre-injected fuel is performed. In other words, the ignition interval Ds is set so that spray guide combustion is possible. In addition to the pressure increase due to the compression action of the piston 3, this spray guide combustion is performed, whereby a further temperature increase in the combustion chamber is obtained. However, a part of the pre-injected fuel is burned by this spray guide combustion, and most of the fuel is burned as “unburned fuel” after the ignition without being used for the combustion by the ignition of the spark plug 5. It will be in the room. This is because in the stratified mixture formed by the pre-injected fuel, the flame cannot be propagated at a portion relatively distant from the electrodes of the spark plug 5 because the air-fuel ratio is high. However, the unburned fuel is exposed to a high temperature atmosphere by burning a part of the pre-injected fuel in the combustion chamber. Therefore, it is expected that at least a part of the unburned fuel is in a state of being reformed to a physical property with improved combustibility by a low-temperature oxidation reaction under a situation where combustion does not occur. However, the remaining unburned pre-injected fuel in the present invention refers to fuel that remains in an unburned state in the combustion chamber after a part of the pre-injected fuel is not subjected to combustion by ignition of the spark plug 5. Therefore, it is not always required that the unburned fuel is in a state showing specific physical properties.

(3) Main injection Next, the main injection timing Tm before the top dead center of the compression stroke when the predetermined first injection interval Di1 has elapsed from the pre-injection timing (the time of Di1-Ds has elapsed from the ignition timing Ts by the spark plug 5) At time Tm), execution of main injection by the in-cylinder injection valve 61 is started. In the internal combustion engine 1, as will be described later, the main injected fuel is subjected to self-ignition or diffusion combustion, and contributes to the engine output. Therefore, the main injection timing Tm is set to a timing (hereinafter referred to as “appropriate injection timing”) at which the engine output obtained by the combustion of the main injection fuel in an amount determined by the engine load or the like is substantially maximized. However, the combustion of the main injection fuel is started by using a flame generated by the ignition of the pre-spray of the pre-injected fuel as a fire type. That is, the main injection timing Tm is set to an appropriate injection timing, and the first injection interval Di1 is set so that the combustion of the main injected fuel is started from the flame generated by the ignition of the pre-spray. By setting the main injection timing Tm and the first injection interval Di1 in this way, the pre-injection timing Tp is inevitably determined. When combustion of the main injection fuel is started, the temperature in the combustion chamber further increases. As a result, the unburned pre-injected fuel and the main injected fuel are self-ignited in the temperature rising field, and these fuels are subjected to diffusion combustion. At this time, when the unburned combustibility of the pre-injected fuel is enhanced as described above, it is expected that the self-ignition of the fuel after the start of the main injection is further promoted.

  Here, FIG. 4 shows the transition of the heat generation rate in the combustion chamber when the basic combustion control according to the present embodiment is performed. In addition, in FIG. 4, transition of the heat release rate corresponding to four different control forms L1-L4 is shown. In these control modes L1 to L4, the pre-injection timing Tp, the pre-injection fuel amount (that is, the pre-injection execution period), the main injection timing Tm, and the ignition timing Ts are the same, but the main injection fuel amount ( That is, the execution period of the main injection is different for each control mode. That is, the main injection fuel amount is L1> L2> L3> L4. That is, FIG. 4 shows changes in the change in the heat generation rate according to the increase or decrease in the main injection fuel amount when the pre-injection timing Tp, the pre-injection fuel amount, the main injection timing Tm, and the ignition timing Ts are the same. Will be.

Here, in FIG. 4, the primary peak of the heat generation rate appears in the portion Z1 surrounded by a dotted line. The primary peak indicates the heat generated by burning the pre-injected fuel by ignition (that is, heat generated by spray guide combustion). At the time when the primary peak of the heat generation rate appears, main injection has not yet been performed, and the combustion chamber is a flame generated by ignition of the pre-injected fuel, and pre-injected fuel not combusted by the ignition Unburned fuel is present. Here, the unburned residue of the pre-injected fuel will be described with reference to FIG. FIG. 5 shows the correlation between the amount of pre-injected fuel and the combustion efficiency of the pre-injected fuel (hereinafter referred to as pre-combustion efficiency) in each of the three combustion conditions L5 to L7 in the pre-injection in the basic combustion control. It is a figure. Specifically, the pre-injection timing Tp and the ignition timing Ts that are combustion conditions are advanced in the order of L5, L6, and L7 in a state where the ignition interval Ds that is an interval between both timings is constant. Note that FIG. 5 shows the above correlation when main injection is not performed and only pre-injection and ignition (that is, only spray guide combustion) are performed.

The pre-combustion efficiency has a relationship represented by the following equation 1 with the remaining unburned ratio of the pre-injected fuel. That is, the higher the pre-combustion efficiency, the lower the unburned rate of the pre-injected fuel.
Unburned rate of pre-injected fuel = 1- Pre-combustion efficiency (Equation 1)
Here, from FIG. 5, when the pre-injection fuel amount is constant, if the pre-injection timing Tp and the ignition timing Ts are advanced (that is, if the first injection interval Di1 is increased), the pre-combustion efficiency decreases, Therefore, it can be found that the unburnt rate tends to be high. Even when the pre-injection fuel amount is changed, the pre-combustion efficiency and the unburned fuel ratio can be controlled to be constant by adjusting the advance amounts of the pre-injection timing Tp and the ignition timing Ts. . As described above, in the basic combustion control according to the present embodiment, the remaining amount rate of the pre-injected fuel is adjusted by adjusting the pre-injected fuel amount, the pre-injection timing Tp, and the ignition timing Ts (that is, the first injection interval Di). Can be controlled.

  Here, returning to FIG. 4, the execution of the main injection is started at a time Tm after the time when the primary peak of the heat generation rate occurs and before the top dead center of the compression stroke. At this time, as described above, the main injected fuel starts to burn using the flame generated by the ignition of the pre-injected fuel as a fire, then self-ignites with the unburned residue of the pre-injected fuel, and further diffuses. It is used for combustion. As a result, a secondary peak that is the maximum peak of the heat generation rate occurs at the time when the top dead center of the compression stroke has passed. Here, in FIG. 4, as the main injection fuel amount increases (that is, as the main injection period becomes longer), the value of the secondary peak of the heat generation rate increases and the generation time of the secondary peak It is late. This means that the combustion period of the main injected fuel becomes longer as the main injected fuel amount increases. From this, it can be inferred that the unburned residue of the main injection fuel and the pre-injection fuel is provided for diffusion combustion or combustion that can be substantially equated with diffusion combustion.

Furthermore, the self-ignition of fuel generated in the basic combustion control according to the present embodiment will be described based on FIG. FIG. 6 shows the pre-injected fuel amount and the main injected fuel amount in the basic combustion control according to the present embodiment while keeping the total injection amount (the total of the pre-injected fuel amount and the main injected fuel amount) in one combustion cycle constant. The transition of the heat generation ratio in the combustion chamber of each of the two forms L8 and L9 in which the ratio is changed is shown. Further, the ratio of the pre-injected fuel amount is higher in the L9 form than in the L8 form. That is, the L9 form has a larger amount of pre-injected fuel than the L8 form, and as a result, the unburned amount of the pre-injected fuel is also increased. In this case, as shown in FIG. 6, in the form of L9, the secondary peak value of the heat generation rate after the top dead center of the compression stroke is larger than in the form of L8. Furthermore, in the form of L9, the falling speed from the secondary peak value of the heat generation rate (the slope of the graph after the secondary peak) is larger than in the form of L8. In the combustion of the pre-injected fuel after the start of the main injection and the combustion of the main injected fuel, the combustion by self-ignition is more promoted in the L9 form than in the L8 form (that is, by self-ignition). It is presumed that this means that the proportion of fuel that burns is high and the proportion of fuel that burns by diffusion combustion is low. From this, it is considered that the unburned residue of the pre-injected fuel contributes to the promotion of self-ignition of the fuel after the main injection. Further, in the basic combustion control according to the present embodiment, the self-ignition of the fuel after the main injection is performed even when the remaining amount of the pre-injected fuel is increased by adjusting the pre-injection timing Tp and the ignition timing Ts in addition to the pre-injected fuel amount. The inventors of the present invention have confirmed that is promoted. That is, in the basic combustion control according to the present embodiment, by adjusting the parameters related to pre-injection and ignition to increase the unburned ratio of the pre-injected fuel, the unburned pre-injected fuel and the main injection after the start of the main injection execution It is possible to promote self-ignition in combustion with fuel.

  As described above, in the basic combustion control according to the present embodiment, the main injection is performed after the spray injection combustion by the pre-injection and the ignition by the spark plug 5, thereby causing the self-ignition and diffusion combustion of the fuel. Let Therefore, it is considered that the combustion by the basic combustion control is similar to so-called diesel combustion or can be substantially identified. Therefore, the air-fuel ratio of the air-fuel mixture in the combustion chamber can be set to an extremely high lean air-fuel ratio (about 20 to 70). Further, in order to realize combustion at such a lean air-fuel ratio, in the combustion control according to the present embodiment, the opening degree of the throttle valve 71 is made larger than that in the conventional gasoline engine combustion control (homogeneous stoichiometric control). . Therefore, the pump loss in the internal combustion engine 1 can be reduced. Furthermore, since the combustion contributing to the engine output is performed by self-ignition and diffusion combustion, the cooling loss in the internal combustion engine 1 can also be reduced as compared with the conventional homogeneous stoichiometric control. Therefore, according to the basic combustion control according to the present embodiment, it is possible to achieve high thermal efficiency that cannot be realized by conventional combustion control of a gasoline engine.

  Further, as described above, the main injection timing is set to an appropriate injection timing at which the engine output of the internal combustion engine 1 is substantially maximized. Therefore, it is possible to cope with an increase in engine load to some extent by increasing the amount of main injection fuel. However, since the main injection is performed when the pressure in the combustion chamber near the top dead center of the compression stroke is very high, the penetration of the fuel spray injected from the in-cylinder injection valve 61 is reduced. That is, the fuel spray injected by the main injection is difficult to diffuse over a wide range. Therefore, if the amount of the main injection fuel is excessively increased, the oxygen present around the spray of the main injection fuel, that is, the amount of oxygen provided for the combustion of the main injection fuel becomes insufficient with respect to the fuel, As a result, the amount of smoke generated may increase. In the basic combustion control according to the present embodiment, it is necessary to cause self-ignition of the fuel after the main injection. However, if the amount of the main injected fuel increases excessively, the temperature in the combustion chamber is increased by the latent heat of vaporization of the main injected fuel. There is also a risk that combustion will become unstable.

  On the other hand, the pre-injection is performed at the pre-injection timing Tp during the compression stroke. Therefore, it is considered that when the pre-injected fuel is burned by ignition by the spark plug 5, it acts to hinder the engine output of the internal combustion engine 1. However, in the combustion by ignition of the pre-injected fuel to the pre-spray, it is only necessary to form a flame that becomes a fire type for the combustion of the main injected fuel. Therefore, as described above, a part of the pre-injected fuel is subjected to combustion by flame propagation caused by ignition. Therefore, the effect | action which prevents the engine output by the spray guide combustion of pre-injection fuel is small. The remaining unburned fuel of the pre-injected fuel that is not used for spray guide combustion is used for self-ignition or diffusion combustion together with the main injected fuel after the main injection, thereby contributing to the engine output. Therefore, increasing the pre-injected fuel amount and increasing the unburned fuel ratio can cope with an increase in engine load.

  Further, at the main injection timing, the unburned residue of the pre-injected fuel is diffused in a wider range than the fuel spray of the main injected fuel in the combustion chamber. Therefore, it is easy to ensure sufficient oxygen when the unburned residue of the pre-injected fuel is subjected to self-ignition or diffusion combustion. Therefore, when the increase in the pre-injected fuel and the increase in the unburned fuel ratio correspond to the increase in engine load, the amount of smoke generated can be suppressed as compared with the case where the main injected fuel amount is increased.

<High load combustion control>
Next, combustion control during high load operation in the internal combustion engine 1 according to the present embodiment will be described. In the internal combustion engine 1 according to this embodiment, it is necessary to increase the fuel injection amount into the combustion chamber as the engine load increases. However, as described above, if the amount of the main injection fuel is excessively increased, the amount of smoke generated increases, and the temperature in the combustion chamber decreases due to the latent heat of vaporization of the fuel, resulting in unstable combustion. There is a risk of becoming. Further, as described above, when increasing the pre-injected fuel amount, the pre-injection timing Tp is advanced together with the increase, that is, the first injection interval Di1 is increased, thereby suppressing the amount of smoke generated. it can. However, there is an upper limit value in the first injection interval Di1 because it is necessary to use the flame generated by ignition of the pre-injected fuel as a fire type for the combustion of the main injected fuel.

  Here, if the pre-injected fuel amount is further increased while the first injection interval Di1 is maintained at the upper limit value, when the main injection is executed, the spray of the main injected fuel is formed in the combustion chamber. The unburned amount of the pre-injected fuel existing at the position increases. As a result, when the unburned portion of the pre-injected fuel and the main injected fuel overlap when the main injection is executed, the fuel concentration in the overlapping portion may become excessively high. In this case, the amount of oxygen present around the portion, that is, the amount of oxygen provided for combustion of the fuel present in the portion becomes insufficient with respect to the fuel, and as a result, smoke is likely to be generated.

  In other words, not only when the main injection fuel amount is increased, but also when the pre-injection fuel amount is excessively increased, there is a risk of increasing the amount of smoke generated. Therefore, in the present embodiment, in a high load region where a relatively large amount of fuel is required to be injected into the combustion chamber during one combustion cycle, instead of the basic combustion control described above, a high amount is required to suppress the amount of smoke generated. Load combustion control is performed. Hereinafter, the high load combustion control according to the present embodiment will be described with reference to FIG. FIG. 7 shows the flow of fuel injection and ignition related to high-load combustion control performed in the internal combustion engine 1 (see the upper part of FIG. 7A) and the fuel injection and ignition in a time series that proceeds from the left side to the right side of the figure. 8 schematically shows the transition of events related to combustion assumed to occur in the combustion chamber (see the lower part of FIG. 7A). FIG. 7B shows a temporal correlation between the port injection, the pre-injection, and the main injection, which are the fuel injections shown in FIG. 7A, and ignition. It should be noted that the form shown in FIG. 7 is schematically shown only for explaining the high-load combustion control according to the present embodiment, and the present invention should not be interpreted as being limited to this form.

  The high-load combustion control according to this embodiment is different from the basic combustion control in the pre-combustion mode. Specifically, in the basic combustion control, as described above, fuel injection for pre-combustion is performed by pre-injection by the in-cylinder injection valve 61, whereas in high-load combustion control, fuel injection for pre-combustion is performed. The port injection by the port injection valve 62 and the pre-injection by the in-cylinder injection valve 61 are performed in combination. Hereinafter, pre-combustion related to basic combustion control may be referred to as “first pre-combustion”, and pre-combustion related to high-load combustion control may be referred to as “second pre-combustion”.

In the high load combustion control according to the present embodiment, as shown in FIG. 7, in one combustion cycle, first, port injection is executed by the port injection valve 62 during the intake stroke. Here, as shown in FIG. 7B, the port injection start timing (hereinafter simply referred to as “port injection timing”) is defined as Tpt. The port injection timing Tpt is set to a timing at which the requested port injection fuel amount can be injected during the intake stroke (that is, during the valve opening period of the intake valve 9). By performing this port injection, an air-fuel mixture is formed in the combustion chamber. Then, similarly to the basic combustion control, pre-injection is performed by the in-cylinder injection valve 61 at the pre-injection timing Tp during the compression stroke. By performing this pre-injection, the pre-spray of the pre-injected fuel injected from the in-cylinder injection valve 61 passes through the ignitable region 5a of the spark plug 5 in the combustion chamber. In this case, a pre-spray is formed in the air-fuel mixture formed by the port injection fuel. As a result, a stratified mixture is formed in the combustion chamber. Then, at the ignition timing Ts in the compression stroke, the pre-spray is ignited by the spark plug 5. By this ignition, spray guide combustion with pre-injected fuel is performed. That is, the ignition interval Ds, which is the interval from the start of execution of pre-injection until ignition is performed, is set so that spray guide combustion is possible, as in the case of basic combustion control.

  Here, in the second pre-combustion, similarly to the first pre-combustion in the basic combustion control, it is only necessary to perform combustion to the extent that a flame that serves as a fire type for the combustion of the main fuel is formed. However, in the second pre-combustion according to the present embodiment, a pre-spray is formed by the pre-injected fuel in the air-fuel mixture formed by the port injected fuel, and the pre-spray is ignited to perform spray guide combustion. Is called. Therefore, it is considered that the port-injected fuel is also burned by flame propagation caused by spray guide combustion. At this time, if the air-fuel ratio of the air-fuel mixture formed by the port-injected fuel is too low, an excessive amount of port-injected fuel may be burned due to flame propagation caused by spray guide combustion with the pre-injected fuel. In this case, the port-injected fuel burned by the second pre-combustion does not contribute to the output of the internal combustion engine 1, so that the thermal efficiency is lowered. Therefore, the air-fuel ratio of the homogeneous air-fuel mixture formed by the port-injected fuel is such that the port-injected fuel combusted by flame propagation generated by spray-guide combustion with the pre-injected fuel is suppressed to a part of the amount of port-injected fuel. Thus, the port injection fuel amount is set.

  Then, execution of the main injection by the in-cylinder injection valve 61 is started at the main injection timing Tm before the compression stroke top dead center when the predetermined first injection interval Di1 has elapsed from the pre-injection timing. Thereby, combustion of the main injection fuel is started using the flame generated by the second pre-combustion as a fire type. When combustion of the main injection fuel is started, the temperature in the combustion chamber further increases. As a result, the port-injected fuel, the unburned pre-injected fuel, and the main-injected fuel are self-ignited in the temperature rising field, and these fuels are subjected to diffusion combustion. Thereby, main combustion is performed. At this time, if the unburned residue of the port injection fuel and the pre-injected fuel existing in the combustion chamber is exposed to a high-temperature atmosphere during the second pre-combustion, the combustibility is enhanced. It is expected that fuel self-ignition will be further promoted.

  As described above, in the high load combustion control according to the present embodiment, the port injection fuel and the pre-injected fuel that have not been burned by the flame propagation generated by the ignition by the spark plug 5 are left after the main fuel injection. It burns by ignition or diffusion combustion. Thereby, in addition to the main injection fuel, the port injection fuel and the pre-injection fuel also contribute to the output of the internal combustion engine 1.

  Then, by performing the second pre-combustion using the port injection in addition to the pre-injection, the pre-injected fuel amount is reduced compared to the case where the first pre-combustion is performed using only the pre-injection as in the basic combustion control. can do. Therefore, by using the port injection and the pre-injection in combination, the amount of smoke generated can be suppressed as compared with the case where the same amount of fuel is injected only by the pre-injection.

Further, the port injection fuel forming the air-fuel mixture is distributed more widely in the combustion chamber than the main injection fuel. Therefore, more oxygen present in the combustion chamber can be used during main combustion. Therefore, it is possible to secure sufficient oxygen necessary for the combustion of the port injection fuel. Furthermore, the port-injected fuel is more widely distributed in the combustion chamber than the unburned residue of the pre-injected fuel. Therefore, the fuel concentration of the portion where the main injection fuel and the port injection fuel overlap is lower than the fuel concentration of the portion where the main injection fuel and the pre-injection fuel (unburned fuel) overlap when the first pre-combustion is performed. . Therefore, even if the main injection fuel and the port injection fuel overlap, it is suppressed that the amount of oxygen provided for the combustion of the fuel existing in the overlapping portion of these fuels becomes insufficient. Therefore, the port-injected fuel is unlikely to cause smoke as compared with either the main-injected fuel or the pre-injected fuel.

  FIG. 8 shows the correlation between the engine load of the internal combustion engine 1 and thermal efficiency (see FIG. 8A) and the engine load of the internal combustion engine 1 and the amount of smoke generated in the basic combustion control and the high load combustion control. It is a figure which shows a correlation (refer FIG.8 (b)). A line L10 in FIG. 8A and a line 12 in FIG. 8B indicate the respective correlations in the high-load combustion control. A line L11 in FIG. 8A and a line 13 in FIG. 8B show the respective correlations in the basic combustion control. As shown in FIG. 8B, in the basic combustion control, when the engine load of the internal combustion engine 1 exceeds a predetermined load Qe0, the amount of smoke generated increases rapidly. And in connection with this, as shown to Fig.8 (a), thermal efficiency will fall. On the other hand, in the high load combustion control, even if the engine load of the internal combustion engine 1 exceeds the predetermined load Qe0, it is possible to suppress an increase in the amount of smoke generated until the engine load reaches a certain level. Therefore, even if the internal combustion engine 1 engine load exceeds the predetermined load Qe0, the thermal efficiency can be improved.

As described above, in the operation region where the engine load is higher than the predetermined load, that is, in the operation region where the total fuel injection amount in one combustion cycle is relatively large, instead of the basic combustion control in which the first pre-combustion is performed, By executing the high-load combustion control that performs the second pre-combustion, diesel combustion can be realized while suppressing the amount of smoke generated.

<Combustion control flow>
Here, the control flow of the combustion control according to the present embodiment will be described with reference to FIG. FIG. 9 is a flowchart showing a control flow of combustion control according to the present embodiment. This control flow is stored in advance in the ECU 20, and is repeatedly executed at predetermined intervals by executing a control program stored in the ECU 20 while the internal combustion engine 1 is operating.

  FIG. 10 shows an example of a control map used for combustion control according to this embodiment. In the upper part (a) of FIG. 10, the correlation between the engine load of the internal combustion engine 1 and the pre-injected fuel amount is indicated by a line L21, and the correlation between the engine load and the main injected fuel amount is indicated by a line L22. A correlation with the injected fuel amount is indicated by a line L23, and a correlation between the engine load and a load corresponding injection amount that is a fuel injection amount corresponding to the engine load is indicated by a line L20. In FIG. 10 (a), S1 indicates a fuel injection amount corresponding to the engine load that is a boundary between the operation region R3 (hereinafter referred to as the low load region R3) and R4 (hereinafter referred to as the medium load region R4). (Hereinafter referred to as a first predetermined amount S1). S2 (> S1) represents the fuel injection amount corresponding to the engine load that becomes the boundary between the medium load region R4 and the operation region R5 (hereinafter referred to as the high load region R5) (hereinafter referred to as the second place). Referred to as quantitative S2). In this embodiment, the basic combustion control is performed when the engine load of the internal combustion engine 1 belongs to the low load region R3 or the medium load region R4, and the engine load of the internal combustion engine 1 belongs to the high load region R5. The above-described high-load combustion control is performed. That is, in the present embodiment, the engine load Qe0 that is the boundary between the medium load region R4 and the high load region R5 corresponds to the “predetermined load” according to the present invention.

  In the lower part (b) of FIG. 10, the correlation between the engine load of the internal combustion engine 1 and the pre-injection timing Tp is indicated by a line L31, and the correlation between the engine load and the ignition timing Ts is indicated by a line L30. A correlation with the main injection timing Tm is indicated by a line L32, and a correlation between the engine load and the port injection timing Tpt is indicated by a line L33. The interval between the line L31 and the line L32 indicates the first injection interval Di1, and the interval between the line L31 and the line L30 indicates the ignition interval Ds. The vertical axis in FIG. 10B represents the crank angle (BTDC: Before Top Dead Center) with reference to the top dead center of the compression stroke, and the larger the value, the earlier the compression stroke. means.

In the control flow according to the present embodiment, first, in S101, the engine load Qe of the internal combustion engine 1 is calculated based on the detection value of the accelerator position sensor 22. As another method, the engine load of the internal combustion engine 1 can be calculated based on the flow rate of air flowing through the intake passage 70, that is, the detected value of the air flow meter 72 and the intake pressure in the intake passage 70. Next, in S102, the load corresponding injection amount S0 is calculated based on the engine load Qe calculated in S101. Specifically, the load corresponding injection amount S0 corresponding to the engine load Qe is calculated using the control map indicated by the line L20 in FIG. In this embodiment, as shown by a line L20, the correlation between the two is recorded on the control map so that the load corresponding injection amount S0 increases as the engine load increases.

  Next, in S103, it is determined whether or not the load corresponding injection amount S0 calculated in S102 is equal to or smaller than a second predetermined amount S2. That is, in S103, it is determined whether the engine load Qe of the internal combustion engine 1 belongs to the low load region R3, the medium load region R4, or the high load region R5. If an affirmative determination is made in S103, that is, if the engine load Qe of the internal combustion engine 1 belongs to the low load region R3 or the medium load region R4, the process of S104 is then executed to perform basic combustion control. .

  In S104, the pre-injection fuel amount Sp, the main injection fuel amount Sm, the pre-injection timing Tp, the main injection timing Tm, and the ignition timing Ts for realizing basic combustion control are determined using the control map shown in FIG. Is done.

  Here, the value of each control parameter for realizing the basic combustion control in the low load region R3 or the medium load region R4 will be described. When the load corresponding injection amount S0 is equal to or less than the first predetermined amount S1, that is, when the engine load of the internal combustion engine 1 belongs to the low load region R3, as shown by a line L21 in FIG. Sp is set to the minimum pre-injection fuel amount Spmin. Here, the minimum pre-injection fuel amount Spmin is a lower limit value of the pre-injection fuel amount that can form a flame as a fire type for starting combustion of the main injection fuel when main injection is executed. Here, if the pre-injected fuel amount Sp increases, combustion due to ignition by the spark plug 5 (that is, spray guide combustion) is likely to be promoted, so there is a possibility that the unburned fuel ratio in the pre-injected fuel may be reduced. By setting the amount Sp to the minimum pre-injection fuel amount Spmin, the unburned rate can be made as high as possible. Therefore, in the low load region R3, by setting the pre-injected fuel amount Sp to the minimum pre-injected fuel amount Spmin, high thermal efficiency can be realized while ensuring stable combustion. Further, since the load corresponding injection amount is relatively small in the low load region R3, even if the increase in the engine load is dealt with by increasing only the main injection fuel amount Sm, the amount of smoke generated increases or the main injection amount increases. It is set as an operating region where the possibility of combustion becoming unstable due to the latent heat of vaporization of fuel is low. Therefore, as shown in FIG. 10 (a), in the low load region R3, the increase in the engine load corresponds to the increase in only the main injection fuel amount Sm, and the pre-injection fuel amount Sp is the minimum pre-injection fuel amount Spmin. It is fixed with.

  Further, the main injection timing Tm is set to an appropriate injection timing before the compression stroke top dead center in order to improve the thermal efficiency of the internal combustion engine 1. In the low load region R3, the first injection interval Di1 in which the thermal efficiency is in a suitable state when the pre-injected fuel amount Sp is the minimum pre-injected fuel amount Spmin with respect to the main injection timing Tm set to the appropriate injection timing. Is set to pre-injection time Tp. As described above, the pre-injection fuel amount Sp is fixed at the minimum pre-injection fuel amount Spmin in the low load region R3. Therefore, in the low load region R3, the first injection interval Di1 is also maintained constant. Therefore, as shown in FIG. 10 (b), in the low load region R3, when the main injection fuel amount Sm varies due to the variation of the engine load, and the main injection timing Tm varies accordingly, the main injection timing The pre-injection timing Tp also varies in conjunction with the variation in Tm.

  As shown in FIG. 10B, the ignition interval Ds that is the interval between the pre-injection timing Tp and the ignition timing Ts is maintained constant. Therefore, in the low load region R3, when the pre-injection timing Tp varies in conjunction with the variation in the main injection timing Tm, the ignition timing Ts also varies in conjunction with the variation.

Further, in the low load region R3, the correlation between the load corresponding injection amount S0 and the main injected fuel amount Sm indicated by the line L22 follows the following formula 2.
Sm = S0−Sp × α (Expression 2)
α: Remaining unburned rate of pre-injected fuel As described above, in the basic combustion control according to the present embodiment, the unburned unburned pre-injected fuel is self-ignited together with the main injected fuel and is used for diffusion combustion, thereby contributing to engine output . Therefore, from the viewpoint of contributing to the engine output, it can be said that the unburned residue of the pre-injected fuel is equivalent to the main injected fuel. Therefore, an appropriate main injection fuel amount Sm is obtained by calculating a coefficient α indicating the unburned ratio of the pre-injected fuel in advance through experiments or the like and calculating the main injection fuel amount Sm according to the above equation 2 in consideration of the coefficient α. Can be requested. Note that the unburned ratio of the pre-injected fuel varies according to the ignition interval Ds and the first injection interval Di1. Therefore, the coefficient α is a value determined based on these. In the low load region R3, since the ignition interval Ds and the first injection interval Di1 are both constant, the coefficient α in the above equation 2 is also a constant value. Further, in the low load region R3, the pre-injected fuel amount Sp is fixed to the minimum pre-injected fuel amount Spmin for the above-described reason, and therefore, Sp = Spmin in the above equation 2. When the amount of fuel combusted by ignition by the spark plug 5 relative to the pre-injected fuel amount (that is, the amount of fuel combusted by the first pre-combustion) is very small, the coefficient α = 1 is set for control. Also good.

  In the middle load region R4, when the increase in the engine load is dealt with by the increase in the main injection fuel amount Sm alone, the amount of smoke generated increases or the combustion does not occur due to the vaporization latent heat of the main injection fuel. It is set as an operation region where there is a high possibility of becoming stable. Therefore, in the medium load region R4, an increase in engine load is dealt with by increasing not only the main injection fuel amount Sm but also the pre-injection fuel amount Sp. Therefore, in the middle load region R4, as indicated by a line L21 in FIG. 10B, the pre-injected fuel amount Sp is increased as the engine load of the internal combustion engine 1 is higher. As a result, the higher the engine load of the internal combustion engine 1, the greater the amount of unburned pre-injected fuel. When the engine load of the internal combustion engine 1 is a predetermined load Qe0 that is the maximum engine load in the medium load region R4, the pre-injected fuel amount Sp becomes the maximum pre-injected fuel amount Spmax.

  In the middle load region R4, as indicated by a line L22 in FIG. 10A, the main injection fuel amount Sm increases as the engine load of the internal combustion engine 1 increases. In the middle load region R4 as well as the low load region R3, the correlation between the load corresponding injection amount S0 and the main injected fuel amount Sm indicated by the line L22 follows the above equation 2. Further, as described above, in the medium load region R4, the pre-injected fuel amount Sp is increased as the engine load increases. Therefore, as shown by a line L22 in FIG. 10A, the increase ratio of the main injection fuel amount Sm in the medium load region R4 (ratio of the increase amount of the main injection fuel amount Sm to the increase amount of the engine load) is It becomes smaller than the increase ratio of the main injected fuel amount Sm in the low load region R3 where the injected fuel amount Sp is fixed. As a result, it is possible to suppress an increase in the amount of smoke generated due to the increase in the amount of main injection fuel and the occurrence of misfire due to an increase in latent heat of vaporization of the main injection fuel.

In the middle load region R4, as shown in FIG. 10B, the pre-injection timing Tp is advanced so that the first injection interval Di1 becomes larger as the engine load of the internal combustion engine 1 is higher. That is, in the intermediate load region R4, the pre-injection timing Tp is advanced more than the advance amount linked to the advance amount of the main injection timing Tm, and the advance amount becomes larger as the engine load is higher. By controlling the pre-injection timing Tp in this way, even if the amount of unburned fuel increases as the pre-injected fuel amount Sp increases, the unburned portion of the pre-injected fuel and the main injected fuel overlap. The fuel concentration can be further reduced. As a result, the amount of smoke generated due to the overlap of these fuels can be suppressed.

  As shown in FIG. 10 (b), also in the middle load region R4, as in the low load region R3, the ignition interval Ds that is the interval between the pre-injection timing Tp and the ignition timing Ts is maintained constant. Therefore, as the engine load increases, when the pre-injection timing Tp is advanced beyond the advance amount linked to the advance amount of the main injection timing Tm, the ignition timing Ts is also advanced to the same extent as the pre-injection timing Tp. The

  Returning to the description of the flowchart shown in FIG. 9, when the values of the respective control parameters for realizing the basic combustion control are determined using the control map shown in FIG. 10 in S104, the process of S105 is executed next. In S105, according to the pre-injection fuel amount Sp, the main injection fuel amount Sm, the pre-injection timing Tp, the main injection timing Tm, and the ignition timing Ts determined in S104, pre-injection and main injection by the in-cylinder injection valve 61, Ignition by the spark plug 5 is executed. Thereby, the basic combustion control according to the present embodiment is realized.

  On the other hand, if a negative determination is made in S103, that is, if the engine load Qe of the internal combustion engine 1 belongs to the high load region R5, the process of S106 is then executed to execute the high load combustion control.

  In S106, the port injection fuel amount Spt, the pre-injection fuel amount Sp, the main injection fuel amount Sm, the port injection timing Tpt, and the pre-injection timing Tp for realizing the high load combustion control using the control map shown in FIG. The main injection timing Tm and the ignition timing Ts are determined.

  Here, the value of each control parameter for realizing the high load combustion control in the high load region R5 will be described. As indicated by lines L21 and L22 in FIG. 10A, in the high load region R5, the pre-injection amount Sp and the main injection amount Sm are fixed to a constant amount. At this time, the pre-injection amount Sp is fixed to an amount smaller by the minimum port injection amount (port injection amount when the engine load of the internal combustion engine 1 is Qe0) than the maximum pre-injection fuel amount Spmax in the medium load region R4. Further, the main injection amount Sm is fixed to the maximum main injection fuel amount Smmax in the medium load region R4. In the high load region R5, an increase in the engine load corresponds to an increase in the main injection fuel amount Sm. Further, the port injection fuel amount Spt at this time is determined by the fact that the air-fuel ratio of the air-fuel mixture formed by the port injection fuel is the port injection fuel combusted by the flame propagation caused by the spray guide combustion by the pre-injection fuel. The air-fuel ratio is set to be suppressed to a part of the air-fuel ratio.

Here, the correlation between the load corresponding injection amount S0 and the port injected fuel amount Spt in the high load region R5 indicated by the line L23 in FIG.
Spt = (S0− (Sp × α ′ + Sm)) / β ′ (Expression 3)
α ′: Remaining unburned ratio of pre-injected fuel β ′: Unburned unburned ratio of port injected fuel α ′ and β ′ are burned of each fuel when the second pre-combustion is performed using both port injection and pre-injection. It is the remaining rate. These unburned residue rates can be obtained based on experiments and the like. As described above, most of the port-injected fuel contributes to the engine output by being subjected to self-ignition or diffusion combustion together with the main-injected fuel. However, part of the port injection fuel combusted in the second pre-combustion does not contribute to the engine output. Therefore, in the high load region R5, according to the above equation 3, not only the pre-injected fuel unburned rate but also the port injected fuel unburned rate (that is, the port injected fuel amount is not burned by ignition with the spark plug 5 but main combustion). Port injection fuel amount Sp considering the ratio of auto-ignition after injection or diffusion combustion)
t is determined. Thereby, an appropriate port injection fuel amount Spt can be obtained. However, as in the case of the coefficient α in the above equation 2, when the amount of fuel combusted by ignition by the spark plug 5 relative to the pre-injected fuel amount (that is, the amount of fuel combusted by the second pre-combustion) is very small In terms of control, the coefficient α ′ may be set to 1. If the amount of fuel combusted by ignition of the pre-spray by the spark plug 5 with respect to the amount of fuel injected into the port (that is, the amount of fuel combusted by the second pre-combustion) is very small, It is good also as (beta) '= 1.

  Further, in the high load region R5, the pre-injection fuel amount Sp and the main injection fuel amount Sm are fixed to a constant amount. Therefore, as shown by lines L31 and L32 in FIG. The time Tm is also fixed at a certain time suitable for each injection amount. Further, the ignition interval Ds is kept constant even in the high load region R5. Therefore, as indicated by a line L30 in FIG. 10B, the ignition timing Ts is fixed at a constant timing. As a result, the remaining amount of the pre-injected fuel in the high load region R5 is also substantially constant. Further, as described above, the port injection timing Tpt is set to a timing at which the required port injection fuel amount Spt can be injected during the intake stroke (that is, during the valve opening period of the intake valve 9). Therefore, as indicated by a line L33 in FIG. 10B, the port injection timing Tpt varies according to the variation of the port injection amount Spt accompanying the variation of the engine load.

  Returning to the description of the flowchart shown in FIG. 9, when the values of the respective control parameters for realizing the high load combustion control are determined in S106, the process of S107 is executed next. In S107, the port injection fuel amount Spt, the pre-injection fuel amount Sp, the main injection fuel amount Sm, the port injection timing Tpt, the pre-injection timing Tp, the main injection timing Tm, and the ignition timing Ts determined in S106. Port injection by the valve 62, pre-injection and main injection by the in-cylinder injection valve 61, and ignition by the spark plug 5 are executed. Thereby, the high load combustion control according to the present embodiment is realized.

  FIG. 10 is merely an example of a control map used for the combustion control according to the present embodiment. The correlation between the engine load of the internal combustion engine and each control parameter in the basic combustion control and the high load combustion control is shown in FIG. It is not restricted to what is shown in FIG.

  The schematic configuration of the internal combustion engine and its intake / exhaust system according to this embodiment is the same as that of the first embodiment. Also in the present embodiment, the basic combustion control similar to that in the first embodiment is performed. However, in this embodiment, the control mode of the high load combustion control executed in the high load region is different from that of the first embodiment.

  Hereinafter, the high load combustion control according to the present embodiment will be described with reference to FIG. FIG. 11 shows the flow of fuel injection and ignition related to high-load combustion control performed in the internal combustion engine 1 (see the upper part of FIG. 11 (a)), and the fuel injection and ignition in the time series from the left to the right in the figure. FIG. 12 schematically shows the transition of events related to combustion assumed to occur in the combustion chamber (see the lower part of FIG. 11A). FIG. 11B shows the temporal correlation between the port injection and the main injection, which are the fuel injections shown in FIG. In addition, the form shown in FIG. 11 is schematically shown in order to explain the high load combustion control according to the present embodiment, and the present invention should not be interpreted as being limited to this form.

The high load combustion control according to this embodiment is different from the high load combustion control according to the first embodiment in the form of the second pre-combustion. Specifically, in the high load combustion control according to the first embodiment, the fuel injection for the second pre-combustion is performed by both the port injection by the port injection valve 62 and the pre-injection by the in-cylinder injection valve 61. In the high load combustion according to the present embodiment, the fuel injection for the second pre-combustion is performed only by the port injection by the port injection valve 62.

  In the high load combustion control according to the present embodiment, as shown in FIG. 11, in one combustion cycle, port injection is first executed by the port injection valve 62 at the port injection timing Tpt during the intake stroke. Here, the port injection timing Tpt can inject the required port injection fuel amount during the intake stroke (that is, during the valve opening period of the intake valve 9), as in the case of the high load combustion control according to the first embodiment. It is set at the time when. By performing this port injection, an air-fuel mixture is formed in the combustion chamber. Then, at the ignition timing Ts in the compression stroke, the mixture is ignited by the spark plug 5. The second pre-combustion is performed by burning a part of the port injection fuel that forms the air-fuel mixture by this ignition.

  Here, even in the second pre-combustion according to the present embodiment, it is only necessary to perform the combustion to such an extent that a flame serving as a fire type for the combustion of the main fuel is formed as in the second pre-combustion according to the first embodiment. However, if the air-fuel ratio of the air-fuel mixture formed by the port injection fuel is too high, the fuel may not be ignited even if ignition is performed by the spark plug 5. In this case, even if the main injection is executed, the main combustion cannot be performed, so that misfire occurs. On the other hand, if the air-fuel ratio of the air-fuel mixture is too low, excessive port injection fuel may be burned due to flame propagation caused by ignition by the spark plug 5. In this case, the port-injected fuel burned by the second pre-combustion does not contribute to the output of the internal combustion engine 1, so that the thermal efficiency is lowered. Therefore, although the air-fuel ratio of the air-fuel mixture formed by the port-injected fuel is ignited by the ignition by the spark plug 5, a flame is generated, but the air-fuel ratio in which only a part of the port-injected fuel burns in the propagation of the flame The port injection fuel amount is set so that

  Then, execution of main injection by the in-cylinder injection valve 61 is started at the main injection timing Tm before the compression stroke top dead center when a predetermined second injection interval Di2 has elapsed from ignition by the spark plug 5. Thereby, combustion of the main injection fuel is started using the flame generated by the second pre-combustion as a fire type. In other words, the second injection interval Di2 is set so that the combustion of the main injected fuel is started from the flame generated by the second pre-combustion. Similarly to the basic combustion control, the main injection timing Tm is set to an appropriate injection timing in the high load combustion control. By setting the main injection timing Tm and the second injection interval Di2 in this way, the ignition timing Ts is inevitably determined.

  When combustion of the main injection fuel is started, the temperature in the combustion chamber further increases. As a result, the port-injected fuel and the main-injected fuel are self-ignited in the temperature rising field, and these fuels are further subjected to diffusion combustion. Thereby, main combustion is performed. At this time, when the port-injected fuel existing in the combustion chamber is exposed to a high-temperature atmosphere during the second pre-combustion, the combustibility is enhanced in the same way as the unburned fuel remaining in the first pre-combustion. It is expected that the self-ignition of the fuel after the start of the main injection is further promoted.

  As described above, in the high-load combustion control, the port injection fuel that remains unburned in the flame propagation generated by the ignition by the spark plug 5 is the main fuel, similar to the unburned pre-injection fuel in the basic combustion control. It burns by self-ignition after injection or diffusion combustion. Thereby, in addition to the main injection fuel, the port injection fuel also contributes to the output of the internal combustion engine 1. Further, as described above, the port injection fuel is less likely to cause smoke than the main injection fuel as well as the pre-injection fuel. Therefore, the amount of smoke generated can be suppressed by performing pre-combustion using port injection.

Therefore, in the operation region where the engine load is higher than the predetermined load, that is, in the operation region where the total fuel injection amount during one combustion cycle is relatively large, the high load combustion control as described above is executed, so that the first embodiment is performed. As in the case of performing the high load combustion control according to the above, diesel combustion can be realized while suppressing the amount of smoke generated.

<Combustion control flow>
Here, the control flow of the combustion control according to the present embodiment will be described with reference to FIG. FIG. 12 is a flowchart showing a control flow of combustion control according to the present embodiment. This control flow is stored in advance in the ECU 20, and is repeatedly executed at predetermined intervals by executing a control program stored in the ECU 20 while the internal combustion engine 1 is operating. In addition, S101-105 in this control flow is the same as each step in the flow shown in FIG. Therefore, steps in which the same processing as that in the flow shown in FIG. 9 is executed are denoted by the same reference numerals and description thereof is omitted.

  FIG. 13 shows an example of a control map used for combustion control according to this embodiment. In the upper part (a) of FIG. 13, as in FIG. 10 (a), the correlation between the engine load of the internal combustion engine 1 and the pre-injected fuel amount is shown by a line L21, and the correlation between the engine load and the main injected fuel quantity is shown by a line L22. The correlation between the engine load and the port injection fuel amount is indicated by a line L23, and the correlation between the engine load and a load corresponding injection amount that is a fuel injection amount corresponding to the engine load is indicated by a line L20. Further, in the lower part (b) of FIG. 13, the correlation between the engine load of the internal combustion engine 1 and the pre-injection timing Tp is indicated by a line L31, and the correlation between the engine load and the ignition timing Ts is indicated by a line, as in FIG. The correlation between the engine load and the main injection timing Tm is indicated by a line L32, and the correlation between the engine load and the port injection timing Tpt is indicated by a line L33. The interval between the line L31 and the line L32 indicates the first injection interval Di1, and the interval between the line L31 and the line L30 indicates the ignition interval Ds. Further, in the lower part (b) of FIG. 13, the interval between the line L30 and the line L32 indicates the second injection interval Di2.

  Also in the present embodiment, as in the first embodiment, the basic combustion control is performed when the engine load of the internal combustion engine 1 belongs to the low load region R3 or the medium load region R4, and high when the engine load belongs to the high load region R5. Load combustion control is performed. That is, also in the present embodiment, the engine load Qe0 serving as the boundary between the medium load region R4 and the high load region R5 corresponds to the “predetermined load” according to the present invention.

  Also in the control flow according to the present embodiment, when an affirmative determination is made in S103, that is, when the engine load Qe of the internal combustion engine 1 belongs to the low load region R3 or the medium load region R4, as in the first embodiment. The basic combustion control is realized by executing the processes of S104 and S105. On the other hand, if a negative determination is made in S103, that is, if the engine load Qe of the internal combustion engine 1 belongs to the high load region R5, the process of S206 is then executed to execute the high load combustion control.

  In S206, the port injection fuel amount Spt, the main injection fuel amount Sm, the port injection timing Tpt, the main injection timing Tm, and the ignition timing Ts for realizing the high load combustion control using the control map shown in FIG. It is determined.

  Here, the value of each control parameter for realizing the high load combustion control in the high load region R5 will be described. As indicated by a line L23 in FIG. 13A, the port injection fuel amount Spt is fixed to a constant amount in the high load region R5. As described above, the port injection fuel amount Spt at this time is determined by the fact that the air-fuel ratio of the air-fuel mixture formed by the port injection fuel is ignited by the ignition of the spark plug 5, and the flame is generated. Is set so that only a part of the port-injected fuel has an air-fuel ratio to burn. In the high load region R5, since the port injection fuel amount Spt is fixed to a constant amount, the port injection timing Tpt is also fixed to a constant time. As described above, the port injection timing Tpt at this time is set to a timing at which the required port injection fuel amount Spt can be injected during the intake stroke (that is, during the valve opening period of the intake valve 9). .

Further, in the high load region R5, an increase in the engine load corresponds to an increase in the main injection fuel amount Sm. The correlation between the load corresponding injection amount S0 and the main injection fuel amount Sm in the high load region R5 indicated by a line L22 in FIG.
Sm = S0−Spt × β (Formula 4)
β: Unburned rate of port injected fuel β is the unburned rate of port injected fuel when second pre-combustion is performed using only port injection. This unburned residue rate can be obtained based on experiments and the like. According to the above equation 4, the main injected fuel amount Sm is determined in consideration of the unburned ratio of the port injected fuel. Thereby, an appropriate main injection fuel amount Sm can be obtained. However, the amount of fuel combusted by ignition of the pre-spray by the spark plug 5 with respect to the port injection fuel amount (that is, the amount of fuel combusted by the second pre-combustion) is the same as the coefficient β ′ in Equation 3 above. If it is very small, the coefficient β = 1 may be set for control.

  Also in the high load region R5, the main injection timing Tm is set to an appropriate injection timing before the top dead center of the compression stroke. Accordingly, as indicated by a line L32 in FIG. 13B, the main injection timing Tm varies according to the variation of the main injection fuel amount Sm accompanying the variation of the engine load. In the high load region R5, as shown in FIG. 13B, the second injection interval Di2 that is the interval between the ignition timing Ts and the main injection timing Tm is maintained constant. Therefore, in the high load region R5, as indicated by a line L30 in FIG. 13B, the ignition timing Ts varies in conjunction with the variation of the main injection timing Tm.

  Returning to the description of the flowchart shown in FIG. 12, when the value of each control parameter for realizing the high load combustion control is determined in S206, the process of S207 is executed next. In S207, in accordance with the port injection fuel amount Spt, the main injection fuel amount Sm, the port injection timing Tpt, the main injection timing Tm, and the ignition timing Ts determined in S206, the port injection by the port injection valve 62 and the in-cylinder injection valve are performed. The main injection by 61 and the ignition by the spark plug 5 are executed. Thereby, the high load combustion control according to the present embodiment is realized.

  FIG. 13 is merely an example of a control map used for the combustion control according to the present embodiment, and the correlation between the engine load of the internal combustion engine and each control parameter in the basic combustion control and the high load combustion control is shown in FIG. It is not restricted to what is shown in FIG.

  In the second embodiment, the second pre-combustion using only port injection is performed in the high load region. However, the second pre-combustion can be performed in a low load region or a medium load region. In this case, it is possible to further suppress the amount of smoke generated in the low load region or the medium load region. However, in the low load region or the medium load region, since the total fuel injection amount during one combustion cycle is relatively small, it is difficult to secure a sufficient amount of port injection fuel, and the mixing formed by the port injection fuel is difficult. In some cases, the air-fuel ratio of the gas becomes excessively high. In this case, the ignitability of the fuel during pre-combustion is reduced, and as a result, misfire is likely to occur. Therefore, from the viewpoint of fuel ignitability during pre-combustion, it is preferable to perform the first pre-combustion using pre-injection in the low load region or the medium load region as in the second embodiment.

In the high load combustion control according to the first and second embodiments, the port injection is executed during the intake stroke (that is, during the valve opening period of the intake valve 9). However, in high-load combustion control, it is only necessary to form an air-fuel mixture in the combustion chamber by executing port injection. Therefore, the port injection for the second pre-combustion may be executed during the exhaust stroke (that is, during the closing period of the intake valve 9). Even in this case, the fuel injected into the intake port 7 by the port injection valve 62 flows into the cylinder 2 together with the intake air in the intake stroke. Thereby, an air-fuel mixture is formed in the combustion chamber.

DESCRIPTION OF SYMBOLS 1 ... Internal combustion engine 2 ... Cylinder 3 ... Piston 5 ... Spark plug 61 ... In-cylinder injection valve 62 ... Port injection valve 7 ... Intake port 8 ... Exhaust port 9 ... Intake valve 10 ・ ・ Exhaust valve 20 ・ ・ ECU
21 ·· Crank position sensor 22 · · Accelerator position sensor 71 · · Throttle valve 72 · · Air flow meter

Claims (3)

An in-cylinder injection valve capable of injecting fuel into the combustion chamber of the internal combustion engine;
A port injection valve capable of injecting fuel into the intake port of the internal combustion engine;
An ignition device whose relative position to the in-cylinder injection valve is determined such that the spray injected from the in-cylinder injection valve passes through the ignitable region and can be directly ignited.
By performing pre-injection by the in-cylinder injection valve during the compression stroke and igniting the pre-spray formed by the pre-injection by the ignition device in an operation region where the engine load of the internal combustion engine is equal to or less than a predetermined load. A first pre-combustion unit that performs a first pre-combustion by burning part of the fuel injected by the pre-injection;
In the operating region where the engine load of the internal combustion engine is higher than the predetermined load, the air-fuel mixture is formed by executing the port injection by the port injection valve during the exhaust stroke or the intake stroke, and the in-cylinder injection valve during the compression stroke A pre-injection is formed in the air-fuel mixture by performing pre-injection, and a part of the fuel injected by the first pre-injection by igniting the pre-spray by the ignition device and the port A second pre-combustion unit that performs a second pre-combustion by burning part of the fuel injected by the injection;
After the occurrence of the first pre-combustion or the second pre-combustion and before the top dead center of the compression stroke, the fuel generated by the first pre-combustion or the second pre-combustion is used as a starting point. By starting execution of main injection by the in-cylinder injection valve at the main injection timing set to start combustion, fuel self-ignition is generated and at least a part of the fuel injected by the main injection A control device for an internal combustion engine, comprising: a main combustion section that performs main combustion by diffusively burning the fuel. An in-cylinder injection valve capable of injecting fuel into the combustion chamber of the internal combustion engine;
A port injection valve capable of injecting fuel into the intake port of the internal combustion engine;
An ignition device whose relative position to the in-cylinder injection valve is determined such that the spray injected from the in-cylinder injection valve passes through the ignitable region and can be directly ignited.
By performing port injection by the port injection valve during the exhaust stroke or intake stroke, an air-fuel mixture is formed, and during the compression stroke, the air-fuel mixture is ignited by the ignition device in the combustion chamber. A pre-combustion section that performs pre-combustion by burning a part of the fuel injected by
After the pre-combustion has occurred and before the compression stroke top dead center, the cylinder is set at the main injection timing set so that the combustion of the injected fuel is started from the flame generated by the pre-combustion. A main combustion section for starting main injection by the inner injection valve to cause self-ignition of the fuel and at least partially diffusively burn the fuel injected by the main injection to perform main combustion Control device for internal combustion engine. The pre-combustion part is a second pre-combustion part, and the pre-combustion is a second pre-combustion;
During the compression stroke, pre-injection is performed by the in-cylinder injection valve, and a part of the fuel injected by the pre-injection is burned by igniting the pre-spray formed by the pre-injection by the ignition device. And further comprising a first pre-combustion unit that performs first pre-combustion,
In an operation region where the engine load of the internal combustion engine is equal to or less than a predetermined load, the first pre-combustion unit performs the first pre-combustion,
In an operation region where the engine load of the internal combustion engine is higher than the predetermined load, the second pre-combustion unit performs the second pre-combustion,
A flame generated by the first pre-combustion or the second pre-combustion at a time after the occurrence of the first pre-combustion or the second pre-combustion and before the compression stroke top dead center. 3. The internal combustion engine according to claim 2, wherein the main combustion is performed by starting execution of the main injection by the in-cylinder injection valve at a main injection timing set so that combustion of the injected fuel is started from the starting point. Control device.

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