JP5561226B2 - Control device for spark ignition gasoline engine - Google Patents

Description

  The present invention relates to a control device for a spark ignition gasoline engine provided with in-cylinder fuel injection means for injecting fuel into the center of a combustion chamber.

  In order to improve the theoretical thermal efficiency of a spark ignition gasoline engine, it is effective to increase its geometric compression ratio. For example, Patent Document 1 describes a spark-ignition direct injection engine having a high compression ratio with a geometric compression ratio set to 14 or higher.

  Further, as described in Patent Document 2, for example, as a technique for achieving both improvement in exhaust emission and improvement in thermal efficiency, it is known that a lean air-fuel mixture is self-ignited and premixed compression self-ignition combustion is performed. Yes. In such an engine that performs premixed compression self-ignition combustion, if the geometric compression ratio is increased, the compression end pressure and the compression end temperature can be increased. Can do.

  Further, Patent Document 3 below is configured to be able to switch from the lean HCCI combustion state to the spark ignition combustion state as the engine load or the rotational speed increases, and to avoid knocking at the time of switching the fuel state. In addition, EGR is performed to recirculate a part of the exhaust gas to the intake system, and the air-fuel ratio is enriched. At this time, the degree of enrichment of the air-fuel ratio is set according to the actual EGR rate estimated by the EGR rate estimating means. A combustion control apparatus for an internal combustion engine configured to correct is disclosed.

JP 2007-292050 A JP 2007-154859 A JP 2009-091994 A

  Incidentally, the spark-ignition gasoline engine having a high compression ratio as described in Patent Document 1 has the advantage that the thermal efficiency can be effectively improved, while the engine operating state is particularly low in the middle range and high. When in the load range, there is a problem that abnormal combustion such as pre-ignition and knocking (spark knock) is likely to occur.

  In addition, in the engine that performs the premixed compression self-ignition combustion, the air-fuel mixture can be appropriately self-ignited in the operation region on the low load side, but the pre-ignition with a severe pressure increase as the engine load increases. It tends to be easy to do. For this reason, an increase in combustion noise and occurrence of abnormal combustion such as knocking may be caused, and the combustion temperature may be excessively increased to increase Raw NOx. Therefore, as disclosed in Patent Documents 2 and 3, in an engine that performs premixed compression self-ignition combustion in a partial load region, normal forced ignition operation is performed on the higher rotation side than the partial load operation region. However, in an engine in which the geometric compression ratio is set to be high in order to stabilize premixed compression self-ignition combustion, the engine disclosed in Patent Document 1 is used in a high load side operation region where forced ignition operation is performed. There was a problem of causing abnormal combustion as well.

  That is, when premixed compression self-ignition combustion is performed in the low load region of the engine, a homogeneous and lean mixture is formed by injecting fuel at a relatively early time from the intake stroke to the compression stroke. The air-fuel mixture can be heated to a high temperature and pressure by the compression action of the piston, and self-ignited near the compression top dead center. On the other hand, when normal forced ignition operation is performed on the higher rotation side than the partial load operation region, if the reactivity of the air-fuel mixture exceeds the ignition threshold before ignition, pre-ignition or knocking, etc. Abnormal combustion will occur.

  Therefore, when the engine shifts from the low load region to the high load region and performs the forced ignition operation, fuel is injected into the cylinder at a high pressure from the late stage of the compression stroke to the early stage of the expansion stroke (high pressure retarded injection). For this purpose, for example, a fuel injection means for injecting fuel into the cylinder at a high pressure of 30 Mpa or more is required. However, if the premixed compression self-ignition combustion is performed in the low load region of the engine using such a high-pressure fuel injection means, the high-pressure fuel is injected into the cylinder during the intake stroke. There are problems such as adhesion and dilution of the oil and inability to efficiently mix the injected fuel and air. In order to prevent this, pressure adjusting means for adjusting the fuel injection pressure of the fuel injection means is provided, and the fuel pressure injected into the cylinder when performing the premixed compression self-ignition combustion in the low load region of the engine is reduced. When it is configured to decrease, there is a problem that the particle size of the fuel increases and smoke is generated, or the emission deteriorates due to the mixing failure of the air-fuel mixture.

  The present invention has been made in view of the circumstances as described above, and can appropriately perform compression self-ignition combustion without causing problems such as oil being diluted in a low load region of the engine or deterioration of emissions. The present invention provides a control device for a spark ignition type gasoline engine that can be performed and can effectively prevent occurrence of abnormal combustion in a high load region of the engine.

In order to solve the above problems, the present invention relates to a port fuel injection means for injecting fuel into an intake port, an in-cylinder fuel injection means for injecting fuel into the center of the combustion chamber, and an exhaust gas derived from the combustion chamber. A spark ignition gasoline engine equipped with an exhaust purification device for purifying harmful components therein, and igniting the mixture at a delayed timing to increase the temperature of the exhaust gas for a predetermined time after the engine is started A control unit that executes an AWS mode that promotes activation of the exhaust purification device by burning and then switches control from the AWS mode to the steady mode is provided, and the control unit controls the engine as a control in the steady mode . In the low load range, fuel is injected into the intake port during the intake stroke by the port fuel injection means to form a leaner and more homogeneous mixture than the stoichiometric air-fuel ratio. Then, this air-fuel mixture is compressed by a piston to become a self-ignited combustion state, and in the high load range of the engine, fuel is burned from the in-cylinder fuel injection means at a fuel pressure of 30 MPa or more from the compression stroke to the initial stage of the expansion stroke. is injected into the room to form a rich air-fuel mixture than the low load range, rapid combustion is allowed Ru control at a predetermined time delayed timing than the compression top dead center and ignited near compression top dead center in this mixture or, alternatively a pre-stage injection by injecting fuel into the combustion chamber in a compression stroke to form a mixture richer than the center portion of the combustion chamber on the outer peripheral portion of the combustion chamber, between the compression stroke late until the expansion stroke initial Fuel from the in-cylinder fuel injection means is divided into at least two times of the latter stage injection that injects the fuel into the combustion chamber at a predetermined time and forms a rich air-fuel mixture at the center of the combustion chamber than when the preceding stage injection is performed. Multi-stage injection Together it is characterized in that to perform the air-fuel mixture either Kano control system Ru is compressed self-ignition combustion by self-ignition of the injected fuel and air (claim 1).

  According to the present invention, when premixed compression self-ignition combustion (hereinafter also referred to as lean HCCI combustion) is performed in a low load region of the engine, fuel is supplied from the port injector of the port fuel injection means to the intake port at a low pressure in the intake stroke. Since it is configured to inject fuel, a homogeneous and lean air-fuel mixture can be formed in the combustion chamber without causing adverse effects such as dilution of oil caused by fuel adhering to the cylinder liner. The air is heated to a high temperature and high pressure by the compression action of the piston, and can be properly self-ignited near the compression top dead center. Moreover, in the high load region of the engine, the fuel can be effectively vaporized and atomized to be appropriately retarded and burned by injecting the fuel into the combustion chamber at a high pressure by the direct injection injector of the in-cylinder fuel injection means. Therefore, there are advantages that high output can be obtained and knock resistance can be effectively improved. Further, from the forced ignition rapid combustion state in the high load region (hereinafter also referred to as the rapid retarded SI combustion state) or the compression self-ignition combustion by the multistage injection (hereinafter also referred to as the multistage CI combustion state) to the lean HCCI combustion state. When shifting quickly, it takes time to adjust the fuel pressure and it becomes difficult to quickly shift the combustion state, or the particle size of the fuel injected into the combustion chamber increases and smoke is generated It is possible to effectively prevent the occurrence of problems such as deterioration of emissions due to poor mixing of the air-fuel mixture and the like.

  In the present invention, preferably, the in-cylinder fuel injection means includes a high-pressure fuel pump driven by the rotation of the engine (claim 2).

  According to this configuration, the injection pressure of the fuel injected from the in-cylinder fuel injection means can be sufficiently increased by utilizing the driving force of the engine, and the engine load is suddenly changed so that the rapid retard from the lean HCCI combustion state. When shifting to the SI combustion state or the multi-stage CI combustion state, there is an advantage that the fuel injection pressure of the in-cylinder fuel injection means can be increased quickly and appropriately.

  In the above-described configuration, more preferably, the fuel pressure of the in-cylinder fuel injection means is increased and maintained in advance during the execution of control for compression self-ignition combustion in a low load region of the engine. 3).

  According to this configuration, when the engine shifts from the low load range to the high load range and shifts from the lean HCCI combustion state to the rapid retarded SI combustion state or the like, control is performed to increase the fuel pressure of the in-cylinder fuel injection means. There is an advantage that the transition to the combustion state can be performed quickly and appropriately.

  In the above configuration, more preferably, the injection pressure of the fuel injected from the in-cylinder fuel injection means into the combustion chamber is increased as the engine speed increases in the high load range of the engine. (Claim 4).

  According to this configuration, it is possible to shorten the fuel injection time in the high engine speed region where the combustion period represented by the crank angle tends to be short, and to sufficiently ensure the vaporization and atomization time. Thereby, the combustibility of the air-fuel mixture can be effectively increased.

  In the above configuration, more preferably, in a spark ignition type gasoline engine provided with an exhaust gas recirculation means for recirculating the exhaust gas derived from the combustion chamber to the intake system, the intake air is supplied via the exhaust gas recirculation means in a high load region of the engine. According to the present invention, the injection pressure of the fuel injected from the in-cylinder fuel injection means is adjusted in accordance with the amount of exhaust gas recirculated to the system.

  According to this configuration, since the engine load is small at the same rotational speed, the amount of exhaust gas recirculated from the exhaust gas recirculation means to the intake system increases, so that the amount of fresh air that tends to decrease relatively increases. In this region, by setting the fuel pressure target value high and injecting fuel into the cylinder at a high fuel pressure, the turbulence in the cylinder can be strengthened and the turbulent energy in the cylinder can be increased. The combustion period of the air-fuel mixture can be effectively shortened and the combustibility can be enhanced in a state where the amount is small.

In the above configuration, more preferably, the port fuel injection means includes a low pressure fuel pump that injects fuel in a fuel tank toward an intake port, and the fuel discharged from the low pressure fuel pump is supplied to the in- cylinder fuel injection means. It is comprised so that it may supply to the high pressure fuel pump of this (Claim 6).

  According to this configuration, since the in-cylinder fuel injection means and the port fuel injection means can be properly used, the lean HCCI combustion state to the rapid retarded SI combustion state or the multistage CI combustion state according to the change state of the engine load. There is an advantage that the transition to the reverse direction and the reverse direction can be performed quickly and appropriately.

  As described above, according to the present invention, it is possible to appropriately perform compression self-ignition combustion without causing problems such as dilution of oil in the low load region of the engine or deterioration of emissions, and the like. It is possible to provide a control device for a spark ignition gasoline engine that can effectively prevent the occurrence of abnormal combustion in a high load range of the engine.

1 is a schematic view showing an overall configuration of a spark ignition type gasoline engine according to an embodiment of the present invention. It is a block diagram which shows the structure of the control system of an engine. It is sectional drawing which expands and shows the combustion chamber of an engine. It is a figure which shows an example of the control map according to the driving | operation area | region of an engine. It is a figure which compares the state of SI combustion by a high pressure retarded injection, and the state of conventional SI combustion. It is a figure which shows the relationship of each parameter related to the presence period of unburned air-fuel mixture. It is a time chart explaining the control content in the driving load area (alpha) of FIG. It is a time chart explaining the control content in the driving load area (beta) of FIG. It is explanatory drawing which shows the control state of the air-fuel | gaseous charge amount corresponding to an engine torque, and an intake / exhaust system. It is explanatory drawing which shows the control state of the fuel-air mixture filling amount corresponding to engine torque, fuel injection, and an ignition system. It is a flowchart of the engine control performed in PCM. It is a characteristic view for setting the injection pressure of fuel in SI mode. It is a time chart explaining another example of the control content in the driving load area (beta) of FIG. It is sectional drawing which shows the modification of the combustion chamber of an engine. It is explanatory drawing which shows typically the fuel injection performed by control of FIG. 13, and the combustion state of a so-called air-fuel mixture as well.

  1 and 2 show an embodiment of a spark ignition gasoline engine according to the present invention. The engine 1 includes a cylinder block 11 provided with a plurality of cylinders 18 (only one is shown in FIG. 1), A cylinder head 12 disposed on the cylinder block 11 and an oil pan 13 disposed on the lower side of the cylinder block 11 for storing lubricating oil are provided. A piston 14 connected to the crankshaft 15 via a connecting rod 142 is fitted in each cylinder 18 so as to be able to reciprocate. The fuel supplied to the engine 1 only needs to have gasoline as a main component, and all of the fuel may be gasoline, or gasoline containing ethanol (ethyl alcohol) or the like. Good.

  A reentrant cavity 20 is formed on the top surface of the piston 14 as shown in an enlarged view in FIG. The cavity 20 is installed so as to face a direct injection injector 67 described later when the piston 14 is positioned near the compression top dead center. A combustion chamber 19 is defined by the cylinder head 12, the cylinder 18, and the piston 14 having the cavity 20.

  The “combustion chamber”, in a narrow sense, refers to the space above the piston 14 at the top dead center, but the combustion chamber 19 in this specification refers to the upper side of the piston 14 regardless of the vertical position of the piston 14. It refers to a combustion chamber in a broad sense consisting of a space formed in Further, the shape of the combustion chamber 19 is not limited to the shape shown in the figure, and for example, the shape of the cavity 20, the top surface shape of the piston 14, the shape of the ceiling portion of the combustion chamber 19, and the like can be appropriately changed. .

  The geometric compression ratio of the engine 1 is set to a relatively high value, for example, in the range of about 14 to 20, for the purpose of improving the theoretical thermal efficiency, stabilizing the premixed compression self-ignition combustion described later, and the like. Yes.

  An intake port 16 and an exhaust port 17 are formed in each cylinder 18 in the cylinder head 12. The intake port 16 and the exhaust port 17 are respectively provided with an intake valve 21 and an exhaust valve 22 that open and close the opening on the combustion chamber 19 side.

  Of the valve systems that drive the intake valve 21 and the exhaust valve 22 respectively, a hydraulically operated variable mechanism that switches the operation mode of the exhaust valve 22 between a normal mode and a special mode, such as a VVL (Variable Valve Lift), is provided on the exhaust side. ) 71 is provided. Although the VVL 71 is not shown in detail in its configuration, the VVL 71 includes two types of cams having different cam profiles including a first cam having one cam peak and a second cam having two cam peaks, and the first and first cams. A lost motion mechanism for selectively transmitting the operating state of either one of the two cams to the exhaust valve 22. In a state where the operation of the first cam is transmitted to the exhaust valve 22, the exhaust valve 22 operates in a normal mode in which the valve is opened only once during the exhaust stroke (see FIG. 8), and the operation of the second cam is performed in the exhaust valve. In the state of transmitting to the exhaust gas 22, the exhaust valve 22 opens during the exhaust stroke and operates in a special mode (exhaust double opening mode) that opens even during the intake stroke (see FIG. 7).

  The VVL 71 is configured to switch the operation mode of the exhaust valve 22 between the normal mode and the special mode according to the operating state of the engine 1. As a mechanism for switching such a mode, an electromagnetically driven valve system that drives the exhaust valve 22 by an electromagnetic actuator can be employed. Further, the internal EGR is not realized only by opening the exhaust valve 22 only twice, but can also be performed by, for example, a double opening mode of intake air in which the intake valve 21 is opened twice. It can also be realized by providing a negative overlap period in which both the intake valve 21 and the exhaust valve 22 are closed in the intake stroke so that the burned gas remains in the cylinder 18.

  As shown in FIG. 2, the exhaust side valve system having the VVL 71 has a VVT (Variable Valve Timing) 72 configured so that the rotational phase of the intake camshaft with respect to the crankshaft 15 can be changed. And a lift amount variable mechanism composed of a CVVL (Continuously Variable Valve Lift) 73 configured so that the lift amount of the intake valve 21 can be continuously changed. The VVT 72 can employ a known structure such as a hydraulic type, an electromagnetic type, or a mechanical type, and illustration of the detailed structure is omitted. Also, the CVVL 73 can adopt various known structures as appropriate, and the detailed structure is not shown. By the VVT 72 and the CVVL 73, the valve opening timing and the valve closing timing of the intake valve 21 and the lift amount can be changed.

  A direct injection injector 67 that directly injects fuel into each cylinder 18 and a port injector 68 that injects fuel into the intake port 16 are attached to the cylinder head 12.

  As shown in an enlarged view in FIG. 3, the direct injection injector 67 is disposed so that its injection hole faces the inside of the combustion chamber 19 from the central portion of the ceiling surface of the combustion chamber 19, and depends on the operating state of the engine 1. An appropriate amount of fuel is directly injected into the center of the combustion chamber 19 at an appropriate timing. In this example, the direct injection injector 67 is a multi-injection type injector having a plurality of injection holes, and is configured to inject fuel so that the fuel spray spreads radially. As shown by the arrows in FIG. 3, at the timing when the piston 14 is positioned near the compression top dead center, a fuel spray injected radially from the direct injection injector 67 in the center of the combustion chamber is formed on the piston top surface. By flowing along the wall surface of the cavity 20, it reaches around the spark plug 25 described later.

  The cavity 20 is formed so that the fuel spray injected at the timing when the piston 14 is positioned near the compression top dead center is contained therein. The combination of the multi-injection type direct injection injector 67 and the cavity 20 is advantageous in that it shortens the time until the spray reaches around the spark plug 25 after fuel injection and shortens the combustion period. It has become.

  The direct injection injector 67 has a high-pressure fuel pump 63 that pressurizes the fuel supplied from the fuel tank 61 via the low-pressure fuel supply system 56 described below to a high pressure of 30 MPa or more, preferably 40 MPa or more, and the high-pressure fuel pump. A high-pressure fuel supply passage 65 in which a high-pressure fuel pumped from 63 is stored and a common rail 64 that supplies the direct-injection injector 67 of each cylinder 18 is disposed is connected. In-cylinder fuel injection means 62 for injecting fuel at a high pressure is configured.

  The high-pressure fuel pump 63 is composed of, for example, a plunger-type pump that is driven via a timing belt that is stretched between a crankshaft driven by the engine 1 and a camshaft. The maximum pressure of the fuel discharged from the high-pressure fuel pump 63 can be appropriately set depending on the performance of the engine 1 and the properties of the fuel. However, if the fuel injection pressure is excessively increased, the driving loss of the engine 1 increases. Therefore, it is usually set to about 120 MPa or less.

  As shown in FIG. 1, the port injector 68 is arranged facing an independent passage communicating with the intake port 16 to the intake port 16, and injects fuel into the intake port 16. One port injector 68 may be provided for one cylinder 18, and if two intake ports 16 are provided for one cylinder 18, the port injector 68 is provided for each of the two intake ports 16. Also good. The format of the port injector 68 is not limited to a specific format, and various types of injectors can be appropriately employed.

  Although not shown in detail in the port injector 68, a low pressure fuel supply passage 55 for supplying fuel in the fuel tank 61 is connected. The low-pressure fuel supply passage 55 is provided with a low-pressure fuel supply system 56 having an electric or engine-driven low-pressure pump and a pressure-regulating regulator. Port fuel injection means 57 for injecting fuel into the intake port 16 at low pressure is configured.

  Further, the low-pressure fuel derived from the low-pressure pump provided in the low-pressure fuel supply system 56 of the port fuel injection means 57 via the regulator is supplied to the high-pressure fuel pump 63 of the in-cylinder fuel injection means 62. The fuel is supplied to the direct injection injector 67 through the common rail 64 and the high-pressure fuel supply passage 65 while being pressurized to a predetermined pressure by the high-pressure fuel pump 63.

  As shown in FIG. 3, a spark plug 25 that ignites the air-fuel mixture in the combustion chamber 19 is attached to the cylinder head 12. The spark plug 25 is disposed through the cylinder head 12 so as to extend obliquely downward from the exhaust side of the engine 1, and the tip of the spark plug 25 is disposed in the center portion of the combustion chamber 19. The direct injection injector 67 is disposed in the vicinity of the tip of the direct injection injector 67 so as to face the combustion chamber 19.

  An intake passage 30 is connected to one side of the engine 1 so as to communicate with the intake port 16 of each cylinder 18. On the other hand, an exhaust passage 40 for discharging burned gas (exhaust gas) from the combustion chamber 19 of each cylinder 18 is connected to the other side of the engine 1.

  An air cleaner 31 that filters intake air is disposed on the upstream side of the intake passage 30. A surge tank 33 is disposed near the downstream end of the intake passage 30. The intake passage 30 downstream of the surge tank 33 is an independent passage branched for each cylinder 18, and the downstream end of each independent passage is connected to the intake port 16 of each cylinder 18.

  Between the air cleaner 31 and the surge tank 33 in the intake passage 30, a water-cooled intercooler / warmer 34 that cools or heats the air and a throttle valve 36 that adjusts the amount of intake air to each cylinder 18 are arranged. Has been. Further, an intercooler bypass passage 35 that bypasses the intercooler / warmer 34 is connected to the intake passage 30, and the intercooler bypass passage 35 is connected to an intercooler for adjusting the flow rate of air passing through the passage 35. A cooler bypass valve 351 is provided. By adjusting the opening degree of the intercooler bypass valve 351, the ratio of the passage flow rate of the intercooler bypass passage 35 and the passage flow rate of the intercooler / warmer 34 is changed, and the temperature of fresh air introduced into the cylinder 18 is changed. Has been adjusted.

  The upstream side portion of the exhaust passage 40 is constituted by an exhaust manifold having an independent passage that is branched for each cylinder 18 and connected to the outer end of the exhaust port 17, and a collecting portion where the independent passages gather. . An exhaust purification device including a direct catalyst 41 and an underfoot catalyst 42 for purifying harmful components in the exhaust gas is provided downstream of the exhaust manifold in the exhaust passage 40. The direct catalyst 41 and the underfoot catalyst 42 include, for example, a cylindrical case and a three-way catalyst disposed therein.

  A portion between the surge tank 33 and the throttle valve 36 in the intake passage 30 and a portion upstream of the direct catalyst 41 in the exhaust passage 40 are connected by an EGR passage 50 that recirculates a part of the exhaust gas to the intake passage 30. ing. The EGR passage 50 includes a main passage 51 in which an EGR cooler 52 for cooling the exhaust gas with engine coolant is disposed, and an EGR cooler bypass passage 53 for bypassing the EGR cooler 52. The main passage 51 is provided with an EGR valve 511 for adjusting the exhaust gas recirculation amount to the intake passage 30, and the EGR cooler bypass passage 53 adjusts the flow rate of the exhaust gas flowing through the EGR cooler bypass passage 53. An EGR cooler bypass valve 531 is provided for the exhaust gas recirculation means 54 for recirculating the exhaust gas led out from the combustion chamber to the intake system.

  The engine 1 configured as described above is controlled by a control unit including a PCM (powertrain control module) 10. The PCM 10 is constituted by a microprocessor having a CPU, a memory, a counter timer group, an interface, and a path connecting these units.

As shown in FIGS. 1 and 2, detection signals from various sensors SW <b> 1 to SW <b> 14 are input to the PCM 10. These various sensors include an air flow sensor SW1 that detects the flow rate of fresh air on the downstream side of the air cleaner 31, an intake air temperature sensor SW2 that detects the temperature of fresh air, and a downstream side of the intercooler / warmer 34. EGR gas that is disposed in the vicinity of the connection portion between the second intake air temperature sensor SW3 that detects the temperature of fresh air after passing the cooler / warmer 34 and the intake passage 30 in the EGR passage 50 and detects the temperature of the external EGR gas. A temperature sensor SW4, an intake port temperature sensor SW5 that is attached to the intake port 16 and detects the temperature of intake air immediately before flowing into the cylinder 18, and an in-cylinder pressure that is attached to the cylinder head 12 and detects the pressure in the cylinder 18 The sensor SW6 and the exhaust passage 40 are arranged in the vicinity of the connection portion of the EGR passage 50 to control the exhaust temperature and the exhaust pressure. An exhaust temperature sensor SW7 and exhaust pressure sensor SW8 and out, the linear O ₂ sensor SW9 for detecting oxygen concentration in exhaust gas is disposed upstream of the direct catalyst 41, the direct catalyst 41 and underfoot catalyst 42 , A lambda O ₂ sensor SW10 that is disposed between and detects the oxygen concentration in the exhaust, a water temperature sensor SW11 that detects the temperature of the engine coolant, a crank angle sensor SW12 that detects the rotation angle of the crankshaft 15, and a vehicle Accelerator opening sensor SW13 for detecting the accelerator opening corresponding to the amount of operation of an accelerator pedal (not shown), intake and exhaust cam angle sensors SW14 and SW15, and common rail 64 of in-cylinder fuel injection means 62 And a fuel pressure sensor SW16 that detects the fuel pressure that is attached and supplied to the direct injector 67. That.

  The PCM 10 performs various calculations based on the detection signals of the sensors to determine the state of the engine 1 and the vehicle, and in response to this, the direct injection injector 67, the port injector 68, the ignition plug 25, and the VVT 72 on the intake valve side. And CVVL 73, exhaust valve side VVL 71, in-cylinder fuel injection means 62, and various valves (throttle valve 36, intercooler bypass valve 351, EGR valve 511, and EGR cooler bypass valve 531) actuators are supplied with control signals. Is.

  FIG. 4 shows a control map of the engine 1 with the crank rotational speed of the engine 1 and the engine load (target torque) as parameters. In a region where the engine load is relatively low in this control map, a homogeneous and lean air-fuel mixture is automatically generated by the compression action of the piston without ignition by the spark plug 25 for the purpose of improving fuel consumption and exhaust emission. It is set to a CI mode region α in which premixed compression self-ignition combustion is performed. On the other hand, an SI mode region β in which the premixed compression self-ignition combustion is stopped and forced ignition using the spark plug 25 is performed is set in a region where the engine load is relatively high.

  As will be described in detail later, in the CI mode region α, the fuel supplied from the low pressure pump and regulator of the port fuel injection means 57 is basically supplied to the port injector at a low pressure of about 0.3 to 0.4 MPa during the intake stroke. A relatively homogeneous lean air-fuel mixture is formed in the combustion chamber 19 by being injected into the intake port 16 from 68, and at the same time, the lean HCCI (Homogeneous-Charge Compression Ignition) is ignited near the compression top dead center. Combustion mode combustion (premixed compression self-ignition combustion) is performed.

  On the other hand, in the SI mode region β, fuel is injected into the cylinder 18 from the direct injection injector 67 of the in-cylinder fuel injection means 62 at a high pressure of 30 MPa or more basically from the late stage of the compression stroke to the early stage of the expansion stroke. As a result, a homogeneous or stratified air-fuel mixture is formed in the combustion chamber 19 and a rapid retarded SI combustion is performed by rapid flame propagation using the forced ignition using the ignition plug 25 in the vicinity of the compression top dead center. It is comprised so that control to perform may be performed.

  As described above, the geometric compression ratio of the engine 1 is set to a high value of 14 or more (for example, 18). By setting the compression ratio to a high value in this way, the compression end temperature and the compression end pressure can be increased. Therefore, in the CI mode region α, it is advantageous for stabilizing the compression self-ignition combustion. On the other hand, since the engine 1 with this high compression ratio is switched to the forced ignition mode (SI mode) in the high load range, the abnormal combustion such as pre-ignition and knocking occurs as the engine load increases particularly in the low speed range. There is an inconvenience that it tends to occur.

  Therefore, in this engine 1, when the engine 1 is in the high load SI mode region β, abnormal combustion is avoided by performing rapid retarded SI combustion in which the fuel injection mode is greatly different from the conventional one. ing. Specifically, this fuel injection mode is greatly retarded from the latter half of the compression stroke to the early stage of the expansion stroke at a fuel pressure that is significantly higher than the conventional pressure, for example, 30 MPa or more, preferably 40 MPa or more. Within the period, fuel is injected into the cylinder 18 by the direct injection injector 67. This characteristic fuel injection mode is hereinafter referred to as “high pressure retarded injection”.

  FIG. 5 shows the heat generation rate (upper stage in FIG. 5) and the reaction of the unburned mixture in the case of rapid retarded SI combustion (solid line) and the conventional SI combustion (indicated by the broken line) in which fuel injection is performed during the intake stroke. FIG. 6 is an explanatory diagram conceptually showing how the degree of progress (lower part in FIG. 5) differs. The load T and the rotational speed Ne are the same, and the fuel injection amount is also the same. However, the fuel injection pressure is set to be significantly higher in the rapid retarded SI combustion than in the conventional SI combustion.

  First, in the conventional SI combustion, the fuel injection P ′ is executed during the intake stroke. In the combustion chamber 19, a relatively homogeneous air-fuel mixture is formed after the fuel injection P 'and before the piston 14 reaches the compression top dead center. In this example, the spark ignition is executed at a timing θig ′ that is considerably late after the compression top dead center, thereby starting combustion by flame propagation. After the start of combustion, as shown by the broken line waveform in the upper part of FIG. 5, the peak of the heat generation rate is reached when a predetermined period has elapsed from the ignition timing θig ', and the combustion is completed at the subsequent time point θend'.

  Here, the period from the start of fuel injection to the end of combustion can be referred to as a period during which an unburned mixture can exist (an unburned mixture period). As shown by the broken line in the lower part of FIG. 5, the conventional SI combustion in which the reaction of the unburned mixture gradually proceeds during the existence period of the unburned mixture has a very long existence period of the unburned mixture. In the meantime, since the reaction of the unburned mixture continues to proceed, the spark ignition occurs when the reactivity of the unburned mixture exceeds the ignition threshold almost simultaneously with the ignition timing θig ′ or earlier. Regardless of this, the unburned mixture may self-ignite and cause pre-ignition (pre-ignition). The ignition timing θig ′ may be set earlier than the timing shown in FIG. 5, but in such a case, the unburned air-fuel mixture is generated during the flame propagation after the spark ignition even if the preignition is avoided. Abnormal combustion with self-ignition, that is, knocking occurs.

  From the above, when the conventional SI combustion is applied in the high load region such as the SI operation region β in the high compression ratio engine as in the present embodiment (that is, the fuel is supplied at a considerably early timing such as during the intake stroke). It is understood that abnormal combustion such as pre-ignition or knocking is unavoidable even if the spark ignition timing θig ′ is adjusted.

  On the other hand, in the rapid retarded SI combustion, as described above, the fuel is injected at a very high injection pressure of 30 MPa or more (for example, 40 MPa) and in a greatly retarded period of the latter stage of the compression stroke (see FIG. 5). Upper P1, P2). Performing such high-pressure and late-time injection (hereinafter referred to as high-pressure retarded injection) shortens the duration of the unburned mixture and avoids abnormal combustion. That is, the duration of the unburned mixture is the period required for fuel injection from the direct injection injector 67 (fuel injection period (A)) and until the combustible mixture is formed around the spark plug 25 after the injection ends. (A) + (B) + (C) obtained by adding the period (air mixture formation period (B)) and the period until the combustion started by ignition ends (combustion period (C)) is there. As will be described below, the high pressure retarded injection shortens the injection period (A), the mixture formation period (B), and the combustion period (C), thereby shortening the existence period of the unburned mixture. it can.

  First, the high injection pressure relatively increases the amount of fuel injected from the direct injection injector 67 per unit time. For this reason, when the fuel injection amount is constant, as shown in the lower part of FIG. 6, the higher the injection pressure, the shorter the period required to inject the injection amount (fuel injection period). Therefore, the high pressure retarded injection in which the injection pressure is set to be significantly higher than the conventional SI combustion contributes to shortening the fuel injection period (A).

  Further, the high injection pressure is advantageous for atomization of the injected fuel spray and makes the flight distance of the fuel spray longer. For this reason, the higher the injection pressure, the shorter the time required for the fuel to evaporate (fuel evaporating time), and the shorter the time required for the spray to reach around the spark plug 25 (spray reaching time). The mixture formation period (B) is a period obtained by adding the fuel evaporation time and the spray arrival time. Therefore, as shown in the lower part of FIG. 6, the mixture formation period (B) increases as the injection pressure increases. Becomes shorter. Therefore, the high-pressure retarded injection with a high injection pressure contributes to shortening the mixture formation period (B).

  Thus, if the fuel injection period (A) and the air-fuel mixture formation period (B) can be shortened by high injection pressure, the fuel injection timing, more precisely, the injection start timing can be delayed accordingly. Against this background, as shown in FIG. 5, the timings of the fuel injections P1, P2 are delayed to the later stage of the compression stroke. And high-pressure injection of fuel at a late timing of the latter stage of the compression stroke leads to an increase in turbulent energy during the combustion period.

  That is, when the fuel injection timing is delayed to the latter stage of the compression stroke, the turbulent energy increases as the fuel injection pressure increases. On the other hand, even if fuel is injected into the combustion chamber 19 at a high injection pressure, if the timing is too early (for example, during the intake stroke), the time until the ignition timing is long, or the compression stroke Due to the compression from the piston 14 inside, the turbulence in the combustion chamber 19 is attenuated. For this reason, when fuel injection is performed at an early timing such as during the intake stroke, the turbulent energy during the combustion period significantly decreases regardless of the injection pressure level.

  The higher the turbulent energy during the combustion period, the shorter the combustion period. Therefore, when the injection timing is the latter half of the compression stroke, the combustion period (C) becomes shorter as the injection pressure is higher, as shown in the lower part of FIG. That is, the high pressure retarded injection in which fuel is injected at a high pressure in the latter stage of the compression stroke contributes to shortening of the combustion period (C).

  In the lower part of FIG. 6, the combustion period (C) when fuel is injected during the intake stroke at a low injection pressure as in the past is indicated by white circles. As is clear from the comparison with the conventional slow combustion period, when the fuel is injected at the latter stage of the compression stroke at a high injection pressure of, for example, 30 MPa or more as in the high pressure retarded injection of the present embodiment, the combustion period ( It can be seen that C) can be greatly shortened.

  Moreover, an injector having a large number of injection holes, such as the direct injection injector 67 of the present embodiment, is advantageous in shortening the combustion period (C) because the turbulent energy is further increased. Further, the combination of the multi-injector type direct injection injector 67 and the cavity 20 provided in the piston 14 mainly causes the fuel spray injected by the fuel injections P1 and P2 in the latter half of the compression stroke to be mainly injected into the cavity 20. Can diffuse quickly within. This also contributes to shortening of the combustion period (C).

  Here, as shown in FIG. 5, as the high-pressure retarded injection, the fuel was injected in two stages (P1, P2) in the latter half of the compression stroke because the fuel atomization was promoted and the turbulent energy was improved. And aimed at each.

  That is, the first fuel injection P1 can ensure a relatively long air-fuel mixture formation period (B), which is advantageous for fuel vaporization atomization. Then, the timing of the second fuel injection P2 can be set to a timing that is further delayed as long as a sufficient mixture formation period (B) is secured by the first fuel injection P1. This is advantageous for improving the turbulent energy in the combustion chamber 19 and contributes to shortening the combustion period (C). In this case, it is desirable that the ratio of the first fuel injection P1 and the second combustion injection P2 is set such that the injection amount of the second fuel injection P2 is larger than the injection amount of the first fuel injection P2. By doing so, the turbulent energy in the combustion chamber 19 is sufficiently increased, which is advantageous for shortening the combustion period (C) and thus avoiding abnormal combustion.

  As described above, the fuel injection period (A), the mixture formation period (B), and the combustion period (C) can be shortened by high-pressure retarded injection in the rapid retarded SI mode. As a result, as shown in FIG. 5, the period from the fuel injection start timing θinj to the combustion end timing θend (the period in which the unburned mixture is present) is compared with the conventional case where fuel is injected during the intake stroke. It can be greatly shortened. And by shortening the said period, even if the compression ratio is high and the load T is high, the air-fuel mixture can be burned out by proper flame propagation without causing abnormal combustion. That is, as shown in the upper part of FIG. 6, in the conventional SI combustion in which the intake stroke injection is performed at a low injection pressure, the reaction progress of the unburned mixture exceeds the ignition threshold as shown by the white circles. In contrast, abnormal combustion occurs, whereas SI combustion using high-pressure retarded injection, that is, rapid retarded SI combustion, suppresses the progress of the reaction of the unburned mixture, as indicated by the black circles, and abnormal combustion. Can be avoided. It should be noted that the ignition timing is set to the same timing for the white circle and the black circle in the upper part of FIG.

  Moreover, in the rapid retarded SI combustion, since the combustion period (C) is greatly shortened, it is assumed that the combustion start timing based on the ignition timing θig is set to a timing delayed to some extent from the compression top dead center. However, combustion does not continue slowly until the expansion stroke has progressed considerably, and a decrease in thermal efficiency and output torque can be avoided. Of course, if the ignition timing θig is further advanced than in the example of FIG. 5, the thermal efficiency and the output torque may be further improved. However, if the ignition timing θig is advanced, knocking is likely to occur. Therefore, the ignition timing θig should be set as far as possible under the restriction that knocking does not occur.

  FIG. 7 is caused by the fuel injection timing and the lift characteristics of the intake and exhaust valves 21 and 22 in the lean HCCI mode combustion control executed in the low load region of the engine 1, that is, the CI mode region α of FIG. It is a figure which shows a heat release rate (J / deg). As shown in the figure, in the lean HCCI mode, the fuel injection timing is set during the intake stroke, and the port injector 68 of the port fuel injection means 57 is supplied to the intake port 16 at a low pressure of about 0.3 to 0.4 MPa. By injecting a relatively small amount of fuel, a homogeneous lean air-fuel mixture is formed in the combustion chamber 19. This lean air-fuel mixture becomes a high temperature and high pressure state due to the compression action of the piston 14, and self-ignition near the compression top dead center causes combustion accompanied by heat generation as shown by the waveform Qa. Become.

  In the lean HCCI mode, the excess air ratio λ, which is a value obtained by dividing the air-fuel ratio (actual air-fuel ratio) of the air-fuel mixture in the combustion chamber 19 by the stoichiometric air-fuel ratio (14.7), extends over the entire combustion chamber 19. It is set to be 2 or more. However, under such a substantially lean and homogeneous air-fuel ratio, misfire may occur unless the in-cylinder temperature is intentionally increased. Therefore, in the lean HCCI mode, the internal EGR that causes the exhaust gas to flow back into the combustion chamber 19 is performed by driving the VVL 71 and opening the exhaust valve 22 during the intake stroke. That is, the exhaust valve 22 is normally opened only in the exhaust stroke (lift curve EX in FIG. 7), but by opening the exhaust valve 22 in the intake stroke based on driving of the VVL 71 (lift curve EX ′), Exhaust gas is caused to flow backward from the exhaust port 17 to the combustion chamber 19. In this way, by causing the high-temperature exhaust gas to flow back (residual) into the combustion chamber 19, the temperature of the combustion chamber 19 can be increased and self-ignition of the air-fuel mixture can be promoted.

  Here, the amount of exhaust gas remaining in the combustion chamber 19 (internal EGR amount) is set to be larger on the low load side and smaller on the high load side. On the other hand, the amount of air (fresh air) introduced into the combustion chamber 19 is set to be smaller on the low load side and larger on the high load side. As a control for that purpose, in the lean HCCI mode, the lift amount of the intake valve 21 is set to gradually increase as the load increases. A lift curve IN indicated by a one-dot chain line in FIG. 7 is a lift curve when the intake valve 21 is in a small lift state. When the load increases from this state, the lift amount of the intake valve 21 is changed to a broken lift curve. The upper limit is set so as to increase gradually. When the lift amount of the intake valve 21 is increased as described above, the intake valve 21 is closed while the closing timing of the intake valve 21 is fixed in the vicinity of the intake bottom dead center (BDC between the intake stroke and the compression stroke). The opening / closing timing and the lift amount of the intake valve 11 are adjusted by the VVT 72 and the CVVL 73 so that only the opening timing of the valve gradually advances toward the exhaust top dead center (TDC between the exhaust stroke and the intake stroke).

  In the lean HCCI mode, the internal EGR based on the restart valve (opening during the intake stroke) of the exhaust valve 22 is executed as described above, so that the exhaust gas recirculation by the exhaust gas recirculation means 54 is stopped. That is, when the opening degree of the EGR valve 511 provided in the EGR passage 50 is set to be fully closed, the recirculation of the exhaust gas from the exhaust passage 40 to the intake passage 30 is stopped.

  In the lean HCCI mode, the excess air ratio λ is set to a significantly lean value of 2 or more as described above. When the air-fuel mixture that is set to be lean as described above is burned, the combustion temperature is greatly lowered, so that the cooling loss can be reduced and the thermal efficiency (fuel consumption) can be improved. Note that when λ ≧ 2, the NOx purification action by the three-way catalyst can hardly be expected, but the amount of NOx (raw NOx amount) generated by combustion when λ ≧ 2 is greatly reduced. Even if a special catalyst (for example, NOx trap catalyst) is not provided in addition to the original catalyst, the amount of NOx contained in the exhaust gas can be suppressed to a sufficiently small value.

  In the SI mode region β where the load of the engine 1 is higher than the CI mode region α, the rapid retarded SI combustion control as shown in FIG. 8 is executed. That is, the fuel is injected from the direct injection injector 67 into the combustion chamber 19 before the compression top dead center (P1, P2), and the forced injection by the spark plug 25 is performed after the fuel injection P1, P2, thereby causing the compression top dead center. Control is performed to burn the air-fuel mixture by flame propagation from the timing past the point.

The total injection amount by the fuel injections P1 and P2 is set to an amount capable of forming an air-fuel mixture having a stoichiometric air-fuel ratio (excess air ratio λ = 1) in the combustion chamber 19. Further, in the SI mode region β in which the rapid retarded SI combustion is performed, the engine load is higher than that in the CI mode region α. Therefore, the fuel injection amount is larger than that in the CI mode region. Therefore, in order to secure a large amount of fresh air commensurate with the increased fuel, the CVVL 73 is driven according to an increase in engine load, and the lift amount of the intake valve 21 is further increased (lift curve IN). In the example of FIG. 8, the lift amount is increased while the lift peak position of the intake valve 21 is fixed. For this reason, as the lift amount increases, the opening timing of the intake valve 21 is advanced within the exhaust stroke, and the closing timing is advanced within the compression stroke. Such a change in the opening / closing timing of the intake valve 21 is advantageous in reducing pumping loss.

  Further, in the rapid retarded SI mode combustion state, external EGR for returning the exhaust gas to the intake passage 30 through the EGR passage 50 is executed. As described above, in the execution region of the rapid retarded SI mode (SI mode region β in FIG. 4), the amount of fresh air required is large, so the external EGR region decreases as the engine load (torque) increases. In the vicinity of the full load range, the external EGR amount is set to 0 in order to secure a larger amount of fresh air.

  The stoichiometric air-fuel ratio (λ = 1) mixture formed on the basis of the fuel injected by the fuel injections P1 and P2 has a predetermined timing (in the example shown in the figure) after a relatively short period from the completion of the injections P1 and P2. Triggered by the spark ignition executed immediately after compression top dead center), combustion starts by flame propagation that is faster than usual, and combustion is performed by the time when the expansion stroke is not so late as shown by the waveform Qb in the figure. Complete.

  As described above, in the SI mode region β, rapid retarded SI combustion (flame propagation combustion based on spark ignition) is performed instead of lean HCCI mode combustion (premixed compression self-ignition combustion by self-ignition). When the combustion in the lean HCCI mode is continued in the SI mode region β where the fuel injection amount is relatively high and the total fuel injection amount is large and the load is lower than this, the occurrence of abnormal combustion or soot This is because there is a high possibility of causing an increase.

  Thus, in the SI mode region β set in the high load region of the engine 1, it is difficult to continue the appropriate CI combustion. Therefore, when the load increases to the SI mode region β, the CI combustion state starts from the CI combustion state. The system switches to SI combustion. However, as described above, in the engine 1 of the present embodiment, the geometric compression ratio is set to a considerably high value of 14 or more in order to reliably perform the CI combustion in the partial load region. Therefore, normal SI combustion, that is, control that injects fuel well before compression top dead center (for example, during the intake stroke) and performs spark ignition near the compression top dead center is executed in the SI mode region β. In such a case, there is a concern that abnormal combustion such as the above-described pre-ignition or knocking in which the unburned air-fuel mixture (end gas) self-ignites during the flame propagation is caused. For this reason, special SI combustion based on the rapid retarded SI mode as shown in FIG. 8 is required.

  The form of fuel injection in the rapid retard SI mode is high-pressure retarded injection by the in-cylinder fuel injection means 62 described above. Therefore, fuel is supplied from the direct injection injector 67 of the in-cylinder fuel injection means 62 into the cylinder 18 with the fuel pressure increased to 30 MPa or more by the high-pressure fuel pump 63 during the retard period from the latter stage of the compression stroke to the early stage of the expansion stroke. Inject directly.

  In addition, as shown in FIG. 8, it replaces with the fuel injection divided | segmented twice into the 1st injection P1 of the front | former stage, and the 2nd injection P2 of the back | latter stage, The said high pressure retarded injection is comprised by one injection (collective injection). May be. Furthermore, low-pressure fuel injection by the port fuel injection means 57 may be added during the intake stroke for the purpose of improving intake charging efficiency. In this intake stroke injection, the intake air charging efficiency is improved by the intake air cooling effect accompanying the fuel injection, which is advantageous in improving the torque.

  Here, as described above, since the fuel injection pressure by the in-cylinder fuel injection unit 62 is extremely high, the fuel is directly injected into the cylinder 18 from the port injector 68 of the in-cylinder fuel injection unit 62 during the intake stroke. Then, a large amount of fuel may adhere to the wall surface (cylinder liner) in the cylinder 18 and cause problems such as oil dilution. However, as described above, the intake stroke injection is performed not from the direct injection injector 67 of the in-cylinder fuel injection means 62 but from the port injector 68 of the port fuel injection means 57 that injects fuel with a relatively low fuel pressure. By configuring to inject the fuel into the inside, problems such as oil dilution can be avoided.

  FIG. 9 shows each parameter of the engine 1 with respect to the load fluctuation in the low speed range, that is, the opening degree (b) of the throttle valve 36, the opening degree (c) of the EGR valve 511, and the closing timing in the exhaust double opening mode ( d), control examples of the valve opening timing (e) of the intake valve 21, the valve closing timing (f) of the intake valve 21, and the lift amount (g) of the intake valve are shown.

  FIG. 9A shows the state in the cylinder 18. This figure shows the configuration of the air-fuel mixture in the cylinder with the horizontal axis representing torque (in other words, engine load) and the vertical axis representing the air-fuel mixture charge amount in the cylinder. As described above, the area on the left side of the figure with a relatively low load is the CI mode, and the area on the right side of the figure with a higher load than the predetermined load is the SI mode. The fuel amount (total fuel amount) is increased as the engine load increases regardless of the CI mode and the SI mode. A fresh air amount for setting the stoichiometric air-fuel ratio (λ = 1) is set with respect to this fuel amount, and this new air amount increases with an increase in the fuel amount as the load increases. It will be.

  In the CI mode in which lean HCCI combustion is performed, as described above, since the internal EGR gas is introduced into the cylinder 18, the remaining amount of the filling amount is constituted by the internal EGR gas and excess fresh air. Therefore, in the CI mode, a lean air / fuel mixture is obtained. On the other hand, in the SI mode in which the rapid retarded SI combustion is performed, the engine 1 is operated so that λ = 1, and the introduction of the internal EGR gas is stopped.

  As shown in FIG. 9B, the throttle valve 36 is set to fully open regardless of the load level of the engine 1 so that the state in the cylinder 18 becomes as shown in FIG. The On the other hand, as shown in FIG. 9C, the EGR valve 511 remains closed in the CI mode, whereas it is opened in the SI mode. The opening degree of the EGR valve 511 is gradually decreased as the engine load increases so that the lower the load and the higher the load in the SI mode. More precisely, it is fully open at the switching between the CI mode and the SI mode, and is fully closed at the fully open load. Therefore, in this control example, even in the SI mode, the introduction of the external EGR gas into the cylinder 18 is stopped at the fully open load.

  FIG. 9D shows the closing timing of the exhaust valve 22 in the exhaust double opening mode. In the CI mode, the valve closing timing is set to a predetermined timing between the exhaust top dead center and the intake bottom dead center in order to introduce the internal EGR gas into the cylinder 18 as described above. On the other hand, in the SI mode, the valve closing timing is set to the exhaust top dead center. That is, in the SI mode, the internal EGR is stopped by stopping the exhaust valve 22 from being opened twice.

  Further, as shown in FIG. 9E, in the CI mode, the valve opening timing of the intake valve 21 is advanced so as to approach the exhaust top dead center as the load of the engine 1 increases. Therefore, the internal EGR gas introduced into the cylinder 18 increases as the load on the engine 1 decreases, whereas the internal EGR gas introduced into the cylinder 18 decreases as the load on the engine 1 increases. . As the load on the engine 1 is lower, the compression end temperature in the cylinder 18 is increased by a large amount of internal EGR gas, which is advantageous in realizing stable compression self-ignition combustion. By suppressing the EGR gas, an increase in the compression end temperature in the cylinder 18 is suppressed, which is advantageous in suppressing premature ignition.

  On the other hand, the opening timing of the intake valve 21 is constant at the exhaust top dead center in the SI mode as shown in FIG. 9E, and the closing timing of the intake valve 21 is as shown in FIG. 9F. Furthermore, the intake bottom dead center is constant in the CI mode and the SI mode. Therefore, in the SI mode, the throttle valve 36 is made constant when fully opened (FIG. 9B), the opening timing and closing timing of the intake valve 21 are made constant, and the lift amount is made constant at the maximum. (FIG. 9 (g)). From this, the ratio between the amount of fresh air introduced into the cylinder 18 and the amount of external EGR gas is adjusted by adjusting the opening of the EGR valve 511. Such control is advantageous in reducing pump loss. Further, in the SI mode, introducing the external EGR gas into the cylinder 18 is advantageous in reducing cooling loss, avoiding abnormal combustion, and suppressing Raw NOx.

  FIG. 10 shows control parameters with respect to fluctuations in the load of the engine 1, and G / F (b), injection timing (c), fuel pressure (d), fuel injection pulse width (e) corresponding to the injection period, And the change of ignition timing (f) is shown.

  Since the state of the air-fuel mixture in the cylinder 18 is as shown in FIG. 10 (a), as shown in FIG. 10 (b), in the CI mode, the G / F gradually increases from the lean as the fuel amount increases. It approaches the air-fuel ratio. In the SI mode, since the external EGR gas is introduced into the cylinder 18, the SI mode gradually decreases as the engine load increases while continuing to the CI mode.

  As shown in FIG. 10C, the fuel injection timing is set, for example, during the intake stroke between the exhaust top dead center and the intake bottom dead center in the CI mode. The fuel injection timing may be changed according to the load of the engine 1. In contrast, in the SI mode, the fuel injection timing is set to the retard period from the latter half of the compression stroke to the early stage of the expansion stroke, whereby high pressure retarded injection is performed. In the SI mode, the injection timing is gradually changed to the retard side as the engine load increases. This is because, as the engine load increases, the pressure and temperature in the cylinder 18 increase and abnormal combustion is likely to occur. To avoid this efficiently, the combustion injection timing is set to the retard side. It is necessary to do.

  Here, the solid line in FIG. 10C shows an example of the fuel injection timing in the case of collective injection in which the high pressure retarded injection is performed by one fuel injection. In contrast, the alternate long and short dash line in FIG. 10C indicates the fuel injection timing of each of the first injection and the second injection when the high pressure retarded injection is divided into two fuel injections of the first injection and the second injection. An example is shown. According to this, since the second injection in the divided injection is executed on the retard side as compared with the case of performing the batch injection, it is advantageous for avoiding abnormal combustion. As described above, this is because the first injection is executed relatively early to secure the vaporization atomization time of the fuel, and the vaporization necessary for relatively reducing the fuel injection amount of the second injection. This is because the atomization time is shortened.

  Further, as indicated by a dotted line in FIG. 10 (c), the overall fuel injection amount increases in the fully open load region of the engine 1, so that the increase in the fuel injection amount is increased for the purpose of improving the intake charging efficiency. Intake stroke injection may be executed.

  FIG. 10D shows a change in the fuel pressure supplied to the direct injection injector 67. In the CI mode, the fuel pressure is set to be constant at the minimum fuel pressure. In contrast, in the SI mode, the fuel pressure is set to be higher than the minimum fuel pressure, and the fuel pressure is set to increase as the engine load increases. This is because abnormal combustion is likely to occur as the engine load increases, and therefore further reduction in the injection period and further delay in the injection timing are required. Further, since the external EGR gas is introduced in the SI mode, particularly when the operating state of the engine 1 is in the middle load range, there is a possibility that the combustion becomes slow and the combustion period becomes long. Therefore, for the purpose of shortening the combustion period, in this control example, the fuel pressure is higher compared to the case where it is set when it is assumed that no external EGR is introduced (see the one-dot chain line in the figure). Adjusted to the value.

  FIG. 10 (e) shows the change in the injection pulse width (injector opening period) corresponding to the injection period in the case of performing batch injection. In the CI mode, the pulse width also increases as the fuel injection amount increases. Similarly, in the SI mode, the pulse width increases as the fuel injection amount increases. However, as shown in FIG. 4D, the fuel pressure in the SI mode is set to be significantly higher than that in the CI mode, so that the fuel injection amount in the SI mode is larger than the fuel injection amount in the CI mode. Nevertheless, the pulse width is set shorter than the pulse width in the CI mode. This shortens the duration of the unburned mixture and is advantageous for avoiding abnormal combustion.

  FIG. 10F shows the change in the ignition timing. In the SI mode, as the fuel injection timing is retarded as the engine load increases, the ignition timing also retards as the engine load increases. Is done. This is advantageous for avoiding abnormal combustion. In the CI mode, although ignition is basically not performed, for the purpose of avoiding smoldering of the spark plug 25, for example, ignition may be performed in the vicinity of the exhaust top dead center as shown by a one-dot chain line in FIG. .

  Next, the control of the engine 1 executed by the PCM 10 will be described in detail with reference to the flowchart shown in FIG. First, after reading the accumulated AWS execution time in step SA1, it is determined in step SA2 whether or not the read AWS execution time is a predetermined value or more. This AWS (Accelerated Warm-up System) is a system that accelerates the purification of exhaust gas by increasing the temperature of the exhaust gas at the start of the engine 1 to accelerate the activation of the catalyst 41, 42. After starting, the control in the AWS mode is executed for a predetermined time. Accordingly, when the determination in step SA2 is NO, that is, when the AWS execution time is not equal to or greater than the predetermined value, the process proceeds to step SA3 to set the AWS mode. In the AWS mode, basically, the amount of intake air is increased and SI combustion is performed in which the ignition timing of the spark plug 25 is significantly retarded.

On the other hand, if YES is determined in step SA2 (i.e., when the AW S execution time is a predetermined value or more) to executes control of the constant mode proceeds to step SA4. In step SA4, after the accelerator opening and the engine speed are read, in step SA5, it is determined whether or not the operating state of the engine 1 is in the CI mode region α in which premixed compression self-ignition combustion (lean HCCI combustion) is performed. When the determination is YES, the process proceeds to step SA6, and the operation mode of the engine 1 is set to the CI mode. On the other hand, when the determination result of step SA5 is NO, the process proceeds to step SA11, and the operation mode of the engine 1 is set to the SI mode.

  In step SA7 in the CI mode, the fuel injection amount injected per cycle is read from a characteristic diagram stored in advance in the PCM 10. This characteristic diagram of the fuel injection amount is set, for example, as a function of the accelerator opening, and is set so that the fuel injection amount increases as the accelerator opening increases. The fuel injection pressure in the CI mode is set to a constant value at a low pressure of, for example, about 0.3 to 0.4 MPa (see FIG. 9D).

  In the subsequent step SA8, the filling amount is controlled. In this filling amount control, as described with reference to FIG. 7, the internal EGR gas is introduced into the cylinder 18 by the control including at least the exhaust double opening mode by the VVL 71. Next, in step SA9, the amount of fuel set in step SSA9 is sucked from the port injector 68 of the port fuel injection means 57 at a predetermined timing in the intake stroke and at a pressure of about 0.3 to 0.4 MPa. Inject into port 16

  Thereafter, in step SA10, the high pressure fuel pump 63 is operated so that fuel can be injected into the cylinder 18 from the direct injection injector 67 of the in-cylinder fuel injection means 62 at a high pressure when the CI mode is shifted to the SI mode. Fuel pressure increase / maintenance control is executed to increase and maintain the pressure of the fuel stored in the common rail 64 in advance to a predetermined value. This increase / maintenance control of the fuel pressure may always be performed in the CI mode. However, for example, it is detected that the engine load tends to increase in the CI mode, and the possibility of shifting from the CI mode to the SI mode is increased. You may comprise so that it may start when it confirms. Further, the fuel pressure in the fuel pressure increase / maintenance control is excessive when the shift from the CI mode to the SI mode is performed, or a value corresponding to the fuel injection pressure set in Step SA16 below, or when the high pressure fuel pump 63 is operated. It is set to a value that does not cause mechanical resistance, for example, about 30 MPa.

  In step SA12 in the SI mode, first, a total injection amount, which means a total amount of fuel injection injected per cycle, is read from a characteristic diagram stored in the PCM 10 in advance. Next, in step SA13, the fuel pressure (target pressure) for the SI mode is read from the characteristic chart stored in the PCM 10 in advance.

  As shown in FIG. 12, the characteristic diagram of the fuel pressure is set as a linear function g for the engine speed and the external EGR rate (the ratio of the external EGR amount to the intake air amount). For example, the target pressure is set to a high value as the engine speed increases in the same load range. Further, at the same rotational speed, the target value of the fuel pressure is increased as the external EGR rate increases as the engine load decreases and the exhaust gas recirculation amount recirculated from the exhaust gas recirculation means 54 to the intake system increases. Set high.

  In step SA14, the combustion injection timing is read based on a preset characteristic diagram (see FIG. 10C), and in step SA15, the ignition timing is read from an ignition map stored in advance in PCM10. This ignition map is a map for setting the ignition timing based on the accelerator opening, and the ignition timing is set to the retard side as the accelerator opening is larger (see FIG. 10F). The ignition timing set here is set to a timing later than the fuel injection timing set in step SA14.

  In this way, when the target fuel pressure, the fuel injection amount and fuel injection timing of the high pressure retarded injection, and the intake stroke injection are executed, the fuel injection amount, the fuel injection timing, and the ignition timing are set, respectively. If the fuel pressure is lower than the target pressure in SA16, the cylinder fuel injection means 62 is controlled to increase the fuel pressure to the target pressure, and then the filling amount control is executed in step SA17. This filling amount control is a control executed to set the air-fuel ratio λ = 1 in accordance with the overall injection amount set in the SI mode operated at the air-fuel ratio λ = 1, and is introduced into the cylinder 18. Control that restricts intake air, control that introduces external EGR gas into the cylinder 18, or control that combines both is executed (see FIG. 9).

  In step SA18, the intake stroke injection of the set fuel injection amount is executed at the set injection timing. In step SA18, fuel is injected into the intake port 16 by the port injector 68 of the port fuel injection means 57 as described above. However, when the fuel injection amount of the intake stroke injection is set to 0, the intake stroke injection in step SA18 is omitted.

  In step SA19, high pressure retarded injection of the set fuel injection amount is executed at the set injection timing. This injection timing is within the retard period from the latter stage of the compression stroke to the expansion stroke process, and the fuel is directly injected into the cylinder 18 by the direct injection injector 67 of the in-cylinder fuel injection means 62. Note that, as described above, the high-pressure retarded injection is performed by split injection including two fuel injections of the first injection and the second injection within the retard period, for example. In step SA23, ignition by the spark plug 25 is performed at the set timing.

  In the spark ignition gasoline engine provided with the port fuel injection means 57 for injecting fuel into the intake port 16 and the in-cylinder fuel injection means 62 for injecting fuel into the center of the combustion chamber 19 as described above, In the low load range, fuel is injected into the intake port 16 by the port fuel injection means 57 in the intake stroke to form a leaner and more homogeneous mixture than the stoichiometric air-fuel ratio, and this mixture is compressed by the piston 14 and compressed. The combustion state to be ignited (lean HCCI combustion state) is set, and in the high load region of the engine 1, the fuel injected into the combustion chamber 19 with a fuel pressure of 30 MPa or more by the in-cylinder fuel injection means 62 is richer than the low load region. An air-fuel mixture is formed, and the air-fuel mixture is ignited in the vicinity of the compression top dead center and burned at a timing delayed by a predetermined period from the compression top dead center (rapid retard S Since the control means (PCM10) is controlled so as to cause combustion (combustion), compression self-ignition is properly performed without causing problems such as oil being diluted in the low load region of the engine 1 or emission deterioration. There is an advantage that the rapid retarded SI combustion can be executed properly while being able to burn and effectively preventing the occurrence of abnormal combustion in the high load region of the engine 1.

  That is, for example, when performing compression self-ignition combustion in the low load region of the engine 1, the fuel is injected into the combustion chamber 19 in the intake stroke by the in-cylinder fuel injection means 62 in which the fuel injection pressure is set to a high pressure of 30 MPa or more. In such a configuration, it is possible to avoid the occurrence of problems such as the oil being diluted by the high pressure sprayed on the cylinder liner and adhering to it, or the oil being diluted or insufficient mixing of the injected fuel and air. I can't. On the other hand, as shown in the above embodiment, when fuel is injected from the port injector 68 of the port fuel injection means 57 to the intake port 16 at a low pressure of about 0.3 to 0.4 MPa in the intake stroke, A homogeneous and lean air-fuel mixture can be formed in the combustion chamber 19 without causing adverse effects such as dilution of the oil caused by the fuel adhering, and this air-fuel mixture is compressed by the compression action of the piston 14. High temperature and high pressure can be generated and self-ignited properly near the compression top dead center.

  Moreover, in the high load region of the engine, the fuel is effectively vaporized and atomized by appropriately injecting the fuel into the combustion chamber 19 by the direct injection injector 67 of the in-cylinder fuel injection means 62, so that the retarded combustion is appropriately performed. Therefore, there is an advantage that high output can be obtained and antiknock property can be effectively improved. Further, when shifting from the rapid retarded SI combustion state in the high load region of the engine 1 to the lean HCCI combustion state in the low load region of the engine 1, the in-cylinder fuel injection means 62 is injected into the combustion chamber 19 in the intake stroke. It is not necessary to execute control for lowering the fuel injection pressure, and it takes a predetermined time to adjust the fuel pressure, so that it is difficult to quickly shift the combustion state or the fuel is injected into the combustion chamber 19. The intake of the fuel from the port injector 68 of the port fuel injection means 57 in the intake stroke without causing problems such as smoke generated due to the increase in the particle size of the fuel or the deterioration of the emission due to the mixing failure of the air-fuel mixture. By injecting fuel into the port 16 at a low pressure, the rapid retarded SI combustion state is rapidly shifted to the lean HCCI combustion state. It is possible.

  Note that, instead of the above-described embodiment in which the high-pressure fuel pump 63 of the in-cylinder fuel injection means 62 is driven by the rotation of the engine 1, a high-pressure fuel pump composed of an electromagnetic pump can be used. However, as shown in the above embodiment, when the high-pressure fuel pump 63 driven by the rotation of the engine 1 is used, the injection pressure of the fuel injected from the in-cylinder fuel injection means 62 is used as the driving force of the engine 1. The fuel injection pressure of the in-cylinder fuel injection means 62 can be quickly and quickly changed when the load of the engine 1 changes suddenly and shifts from the lean HCCI combustion state to the rapid retarded SI combustion state. There is an advantage that it can be increased appropriately.

  Further, as shown in the above-described embodiment, when the lean HCCI combustion control in the low load region of the engine 1 is executed, the fuel injection pressure of the in-cylinder fuel injection means 62 is increased and maintained in advance to a predetermined value. Is a response delay caused by increasing the fuel pressure of the in-cylinder fuel injection means 62 when the engine 1 shifts from the low load range to the high load range and shifts from the lean HCCI combustion state to the rapid retarded SI combustion state. Is effectively prevented and the transition to the rapid retarded SI combustion state can be performed quickly and appropriately.

  The fuel pressure of the in-cylinder fuel injection means 62 that is maintained when the lean HCCI combustion is controlled is the fuel injection pressure that is set when the CI mode lean HCCI combustion state shifts to the SI mode rapid retarded SI combustion state. Or a value that does not cause excessive mechanical resistance when the high-pressure fuel pump 63 is operated, for example, about 30 MPa is desirable. As described above, when the fuel pressure of the in-cylinder fuel injection means 62 is set to a value corresponding to the fuel injection pressure set at the time of transition to the rapid retard SI combustion state, the transition to the rapid retard SI combustion state is performed. There is an advantage that can be executed immediately. On the other hand, when the fuel pressure of the in-cylinder fuel injection means 62 is set to a value that can prevent excessive mechanical resistance when the high-pressure fuel pump 63 is operated, the engine 1 is used to drive the fuel pressure pump 63. Therefore, there is an advantage that it is possible to effectively prevent the deterioration of the fuel consumption by suppressing the necessary driving force from increasing.

  In the above embodiment, when rapid retarded SI combustion is performed in the high load region of the engine 1, the fuel injected from the in-cylinder fuel injection means 62 into the combustion chamber 19 as the engine speed increases. The fuel injection time is shortened in the high rotation region of the engine 1 where the combustion period represented by the crank angle tends to be shortened, and the vaporization and atomization time is sufficiently increased. As a result, the combustibility of the air-fuel mixture can be effectively enhanced.

  Further, the fuel injected from the in-cylinder fuel injection means 62 according to the exhaust gas recirculation amount recirculated to the intake system comprising the intake air passage 30 via the EGR passage 50 of the exhaust gas recirculation means 54 in the high load region of the engine 1. By adjusting the injection pressure of the exhaust gas, the amount of exhaust gas recirculated from the exhaust gas recirculation means 54 to the intake system increases because the engine load is small at the same rotation speed, so that the amount of fresh air is relatively reduced. It is desirable that the air-fuel mixture be sufficiently combustible in the operating region where the air-fuel ratio tends to occur. That is, the target value of the fuel pressure is set to a relatively high value compared to the case where it is assumed that the external EGR is not introduced in the operation region where the fresh air amount tends to decrease. By setting and injecting fuel into the cylinder 18 at a high fuel pressure, the turbulence in the cylinder can be strengthened and the turbulence energy in the cylinder 18 can be increased. The combustion period can be effectively shortened and the combustibility can be enhanced.

  Further, as in the above embodiment, the fuel in the fuel tank 61 is injected toward the intake port 16 by the low pressure fuel pump provided in the low pressure fuel supply system 56 of the port fuel injection means 57 and discharged from the low pressure fuel pump. When the fuel is supplied to the high-pressure fuel pump 63 of the in-cylinder fuel injection means 62, the in-cylinder fuel injection means 62 and the port fuel injection means 57 are properly used, so that the engine load There is an advantage that the transition from the lean HCCI combustion state to the rapid retarded SI combustion state and the transition in the opposite direction can be performed promptly and appropriately in accordance with the change state.

  It should be noted that control in the multi-stage CI mode as shown in FIG. 13 is executed in place of the above-described embodiment configured to cause the rapid retarded SI combustion state in the high load region (SI mode region in FIG. 4) β of the engine 1. You may comprise. That is, in this multi-stage CI mode, fuel is burned from the direct injection injector 67 in two injection timings (Pa, Pb) set in the vicinity of the compression top dead center and a predetermined time in the compression stroke before that. By injecting fuel into the chamber 19, control is performed to combust the air-fuel mixture based on each fuel by self-ignition. In the following description, the first fuel injection Pa executed during the compression stroke is performed at the front stage, and the second fuel injection Pa executed near the compression top dead center after that (in the illustrated example, very early in the expansion stroke). The fuel injection Pb is referred to as post-injection.

  Specifically, in this embodiment, the timing of the front injection Pa in the multi-stage CI mode is, for example, before the top dead center (BTDC) with reference to the compression top dead center (TDC between the compression stroke and the expansion stroke). ) It is set within a period of about 60 to 50 ° CA (CA represents a crank angle), and the timing of the post-injection Pb is set within a period of about 0 to 10 ° CA after top dead center (ATDC). Moreover, the ratio of each injection amount by the front | former stage injection Pa and the back | latter stage injection Pb is set to about 3: 7-7: 3.

  Here, since the operation region β in which the multistage CI mode is executed has a higher load than the lean HCCI mode execution region (CI mode region) α, the total injection amount by the front injection Pa and the rear injection Pa is More than the lean HCCI mode (fuel injection P0 in FIG. 7). The lift amount of the intake valve 11 is set to be larger than that in the lean HCCI mode (lift curve IN), and a relatively large amount of air (fresh air) is introduced into the combustion chamber 19. And, as shown in the waveform Qc of FIG. 13, the mixture of fuel and air that has been separately injected as described above self-ignites near the compression top dead center, so that it has two peaks at different times. Combustion accompanied by heat generation will occur. It should be noted that such a shape of the waveform Qc is merely conceptual, and there are naturally cases where two peaks do not appear clearly.

  As described above, the fuel is injected separately into the front injection Pa and the rear injection Pb in consideration of problems such as combustion noise. That is, in the high load region β of the engine 1 having a large amount of fuel injection, if fuel is injected at a time, a sudden combustion occurs in which a large amount of injected fuel is burnt in a short time, thereby causing There is a risk that the internal pressure will rise rapidly and the combustion noise will increase significantly. Therefore, by dividing and injecting fuel as described above, relatively mild combustion is continuously caused to avoid an increase in combustion noise as described above.

  However, even if the fuel injection is divided into a plurality of times, depending on the arrangement of the direct injection injector 67 and the shape of the piston 14, the fuels injected at each time may be mixed and the mixed fuel burns almost simultaneously. There is. In this way, if combustion occurs in a mixture of fuels with different injection timings, not only does combustion noise become excessive, but the oxygen required for combustion is significantly insufficient locally, resulting in a large amount of soot (carbonaceous matter). Particles) may occur.

  In order to deal with such a problem, in this embodiment, the direct injection injector 67 is disposed at the center of the ceiling portion of the combustion chamber 19, and the cavity 58 positioned on the crown surface of the piston 14 is special as shown in FIG. By being formed into a shape, the separately injected fuel does not burn at the same time, and an increase in combustion noise and a large amount of soot can be avoided as will be described later. The cavity 58 formed at the center of the crown surface of the piston 14 has an upward opening 58a facing the direct injection injector 67 at the upper end, and the area (opening area) of the opening 58a is the cavity. 58 is set to be smaller than the maximum cross-sectional area inside 58 (the maximum value of the horizontal cross-sectional area at each height position of the cavity 58). That is, the cavity 58 is formed in a constricted shape so that the inner diameter becomes narrower toward the upper side in the range from the opening 58a to a predetermined depth.

  An annular recess 59 having an annular shape in plan view is provided on the crown surface of the piston 14 located radially outside the cavity 58 so as to surround the cavity 58. The annular recess 59 is formed such that the height decreases toward the outer side in the radial direction, and the maximum depth (the depth of the outermost peripheral portion) is set to be shallower than the depth of the cavity 58.

  Further, an annular standing wall portion 60 protruding upward from the annular recess 59 is provided on the outermost peripheral portion of the piston 14 positioned further radially outward than the annular recess 59. The protruding height of the standing wall portion 60 is set to be the same as the portion (lip portion) surrounding the opening 58 a at the upper end of the cavity 58.

  In the multi-stage CI mode, in addition to the fuel split injection control as described above, the VVL 71 is driven so as to invalidate the function of depressing the exhaust valve 22 during the intake stroke, and the exhaust valve 22 is opened during the intake stroke. The valve is stopped. As a result, the exhaust gas hardly flows back into the combustion chamber 19 and internal EGR is prohibited. In the multistage CI mode, external EGR is executed instead of the internal EGR. That is, the control to recirculate the exhaust gas from the exhaust passage 40 to the intake passage 30 is executed by opening the EGR valve 511 provided in the EGR passage 50 of the exhaust gas recirculation means 54 to a predetermined opening degree.

  The reason for switching from the internal EGR to the external EGR in this way is to prevent excessive combustion of the combustion chamber 19 and avoid abnormal combustion. That is, the high load region β in which the multi-stage CI mode is executed has a higher engine load than the lean HCCI mode execution region (CI mode region α) and a large amount of total fuel is injected. The amount of heat increases, and the combustion chamber 19 tends to increase in temperature. For this reason, when the internal EGR is continued in the high load region β, the combustion chamber 19 becomes increasingly hot and abnormal combustion such as pre-ignition or knocking may occur. Therefore, by switching from the internal EGR to the external EGR, the exhaust gas cooled after passing through the EGR passage 50 with the EGR cooler 52 is returned to the intake passage 30 to prevent the combustion chamber 19 from being excessively heated. The abnormal combustion as described above is avoided. If the engine load increases, the amount of fresh air required increases accordingly, so the amount of exhaust gas recirculated by the external EGR (external EGR amount) decreases as the load increases.

  Here, the combustion mode realized by the control based on the multi-stage CI mode as described above will be described more specifically with reference to FIGS. FIG. 15A shows a state where the front injection Pa is performed from the direct injection injector 67. As described above, the piston 14 at this time is located at about 60 to 50 ° CA before compression top dead center (BTDC). When fuel is injected radially from a plurality (12) of nozzle holes provided at the tip of the direct injection injector 67 toward the crown surface of the piston 14 at such a position, the spray of the fuel is injected into the piston 14. It goes to the annular recess 59 provided on the outer side in the radial direction of the crown surface.

  The fuel (spray) injected toward the annular recess 59 of the piston 14 is then dispersed while being guided upward by the standing wall portion 42 provided on the outermost peripheral portion of the piston 14, and based on the dispersed fuel, At the piston position (timing before compression top dead center) as shown in FIG. 15B, the air-fuel mixture X1 is formed in the outer peripheral portion of the combustion chamber 19 (mainly inside the annular recess 59 and its upper space). The air-fuel ratio of the air-fuel mixture X1 formed here is set to a theoretical air-fuel ratio (excess air ratio λ = 1) as a local air-fuel ratio only in the outer peripheral portion of the combustion chamber 19. That is, before the compression top dead center, the injection timing and the injection amount of the pre-stage injection Pa are set so that the air-fuel mixture X1 having a concentration about the stoichiometric air-fuel ratio is locally formed on the outer peripheral portion of the combustion chamber 19. Yes.

  Of course, the pre-injection Pa may cause a small amount of fuel to exist in a region other than the outer peripheral portion of the combustion chamber 19 (for example, inside the cavity 58), but the concentration of the fuel is extremely high compared to the outer peripheral portion of the combustion chamber 19. It is thin. In other words, the air-fuel mixture X1 richer than the inside of the cavity 58 is formed in the outer peripheral portion of the combustion chamber 19 by the preceding injection Pa.

  The air-fuel mixture X1 formed on the outer peripheral portion of the combustion chamber 19 as described above is compressed by the rise of the piston 14 to become high temperature and pressure, and when the piston 14 reaches the vicinity of the compression top dead center, FIG. It burns by self-ignition as shown in (CI combustion). In the figure, the region where the air-fuel mixture X1 is burning is shown in black or gray. The region Y1 where the air-fuel mixture X1 burns is limited to the outer peripheral portion of the combustion chamber 19 corresponding to the region where the air-fuel mixture X1 is formed.

  When the combustion based on the front injection Pa as described above starts, the rear injection Pb as shown in FIG. 15 (d) is executed around that time (in the example shown, almost simultaneously with the start of combustion based on the front injection Pa). . As described above, the timing of the post-injection Pb is substantially the same as the time when the piston 14 reaches its rising end (compression top dead center) or about ATDC 0 to 10 ° CA immediately after that. When the fuel is injected from the direct injection injector 67 at such timing (near the compression top dead center), the spray of the fuel is directed toward the inside of the cavity 58 provided at the center of the crown surface of the piston 14. Become. Then, the fuel (spray) injected toward the inside of the cavity 58 is dispersed while being guided upward along the peripheral wall of the cavity 58, and based on the dispersed fuel, as shown in FIG. The air-fuel mixture X2 is formed at the center of the combustion chamber 19 (mainly inside the cavity 58). The local air-fuel ratio of the air-fuel mixture X2 is also set to about the stoichiometric air-fuel ratio (excess air ratio λ = 1), similar to the air-fuel mixture X1 based on the preceding injection Pa. In other words, the latter-stage injection Pb forms the inside of the cavity 58 in the cavity 58 richer than when the first-stage injection Pa is executed (more specifically, based on the fuel injected by the first-stage injection Pa. A mixture X2 (richer than an extremely thin mixture) is formed.

  However, the air-fuel mixture X2 needs to be present before the combustion of the air-fuel mixture X1 based on at least the preceding stage injection Pa ends. That is, before the combustion based on the front stage injection Pa is finished, the mixture X2 having a concentration about the stoichiometric air-fuel ratio is locally formed in the central portion of the combustion chamber 19 so that the rear stage injection Pb is Injection timing and injection amount are set. In this embodiment, the post-injection Pb is executed almost simultaneously with the combustion start timing based on the pre-injection Pa, but the fuel injection pressure from the direct injection injector 67 is set to a fairly high value of 30 MPa or more. Therefore, even if the post injection Pb is executed at a late timing as described above, the air-fuel mixture X2 can be formed before the end of the combustion.

  Here, as described above, both the local air-fuel ratio of the air-fuel mixture X2 based on the post-injection Pb and the local air-fuel ratio of the air-fuel mixture X1 based on the front-stage injection Pa executed before this are both Since the air-fuel ratio is approximately the stoichiometric air-fuel ratio and these air-fuel mixtures X1 and X2 are formed separately in the combustion chamber 19, the average excess air ratio λ of the entire combustion chamber is set to a value greater than 1. It will be.

  Based on the post-injection Pb as described above, the air-fuel mixture X2 is formed near the compression top dead center and while the combustion by the pre-injection Pa is continuing (before the end of the combustion). As shown in (f), after the post-injection Pb, self-ignition occurs and burns in a short time. The region Y2 where the air-fuel mixture X2 burns is limited to the central portion of the combustion chamber 19 corresponding to the region where the air-fuel mixture X2 is formed. That is, the air-fuel mixture X1 based on the upstream injection Pa described above burns in the outer peripheral portion (fuel region Y1) of the combustion chamber 19 corresponding to the installation portion of the annular recess 59, whereas the air-fuel mixture X2 based on the rear injection Pb Then, combustion occurs in the central portion of the combustion chamber 19 corresponding to the installation portion of the cavity 58 (the combustion region Y2 located closer to the radial center than the fuel region Y1).

  As described above, in the multi-stage CI mode, a relatively large amount of fuel corresponding to the load T is injected in a plurality of times (pre-stage injection Pa and post-stage injection Pb), so that the air-fuel mixture (X1, X2) is divided into separate spaces. ) And self-ignite and burn them independently. In the high load region β in which such a multi-stage CI mode is executed, the separately injected fuels are not mixed and burned at the same time. There is no worry of increasing soot due to a lack of oxygen. In addition, the air-fuel mixtures X1 and X2 based on the front injection Pa and the rear injection Pb are locally set to have an excess air ratio of about λ = 1, so that the exhaust gas generated by combustion in such an environment is used. If present, there is an advantage that harmful components can be sufficiently purified only by the three-way catalyst.

1 engine 10 PCM (control means)
14 piston 16 intake port 19 combustion chamber 25 spark plug 50 EGR passage 54 exhaust gas recirculation means 57 port fuel injection means 62 in-cylinder fuel injection means 67 direct injection injector 68 port injector

Claims (6)

Port fuel injection means for injecting fuel into the intake port, in-cylinder fuel injection means for injecting fuel into the center of the combustion chamber, and an exhaust purification device for purifying harmful components in the exhaust gas derived from the combustion chamber A spark ignition gasoline engine,
For a predetermined time after the engine is started, the AWS mode is executed to promote the activation of the exhaust purification device by igniting and burning the air-fuel mixture at a delayed timing in order to increase the temperature of the exhaust gas, and then the AWS mode Control means for switching the control from the steady mode to
As the control in the steady mode, the control means,
In the low engine load range, fuel is injected into the intake port during the intake stroke by the port fuel injection means to form a leaner and more homogeneous mixture than the stoichiometric air-fuel ratio, and this mixture is compressed by the piston and self-ignited. Let the combustion state
In the high load region of the engine, fuel is injected into the combustion chamber from the in-cylinder fuel injection means at a fuel pressure of 30 MPa or more from the compression stroke to the initial stage of the expansion stroke to form a richer air-fuel mixture than in the low load region. , the combustion chamber air-fuel mixture into or rapid combustion is allowed Ru control at a timing delayed a predetermined time period than the compression top dead center and ignited near compression top dead center, or by injecting fuel into the combustion chamber in the compression stroke Fuel is injected into the combustion chamber at a predetermined timing between the front stage injection that forms a richer air-fuel mixture at the outer peripheral part than the center part of the combustion chamber, and the late stage of the compression stroke to the early stage of the expansion stroke, and The fuel is injected in multiple stages from the in-cylinder fuel injection means in at least two times with the latter stage injection that forms a richer mixture than when the former stage injection is performed, and the mixture of the injected fuel and air is Self ignition and compression self Control apparatus for a spark ignition gasoline engine and executes either Kano control system Ru is fire burning. In the control device of the spark ignition type gasoline engine according to claim 1,
The in-cylinder fuel injection means includes a high-pressure fuel pump driven by the rotation of the engine. In the control device of the spark ignition type gasoline engine according to claim 1 or 2,
A control apparatus for a spark ignition gasoline engine, characterized in that the fuel pressure of the in-cylinder fuel injection means is increased and maintained in advance during the execution of control for compression self-ignition combustion in a low load region of the engine. In the control apparatus of the spark ignition type gasoline engine according to any one of claims 1 to 3,
A spark-ignited gasoline configured to increase the injection pressure of fuel injected from the in-cylinder fuel injection means into the combustion chamber as the engine speed increases in a high load range of the engine. Engine control device. In the control device of the spark ignition type gasoline engine according to any one of claims 1 to 4,
Provided with exhaust gas recirculation means for recirculating exhaust gas derived from the combustion chamber to the intake system;
The fuel injection pressure of the in-cylinder fuel injection means is adjusted in accordance with the exhaust gas recirculation amount recirculated to the intake system via the exhaust gas recirculation means in a high load region of the engine. A spark ignition gasoline engine control device. In the control device of the spark ignition type gasoline engine according to any one of claims 1 to 5,
The port fuel injection means includes a low pressure fuel pump that injects fuel in a fuel tank toward an intake port,
A control device for a spark ignition gasoline engine, characterized in that the fuel discharged from the low pressure fuel pump is supplied to the high pressure fuel pump of the in-cylinder fuel injection means.