JP6583336B2 - Control device for compression self-ignition engine - Google Patents

Description

  The technology disclosed herein relates to a control device for a compression self-ignition engine.

  Patent Document 1 describes an engine configured to introduce internal EGR gas into a combustion chamber in an operation region in which an air-fuel mixture in the combustion chamber burns by self-ignition. This engine is configured such that a valve operating mechanism that opens an exhaust valve operates by switching between a normal mode and a special mode.

  Specifically, the normal mode is a mode in which the exhaust valve is opened in the exhaust stroke, and the special mode is a mode in which the exhaust valve is opened in the exhaust stroke and the intake stroke. The valve operating mechanism has a first cam and a second cam. The first cam is a cam selected in the normal mode, and has one cam crest. The second cam is a cam selected in the special mode and has two cam peaks.

  More specifically, after the exhaust valve is opened, the cam profile of the second cam increases gradually and reaches the maximum lift, and then the lift gradually decreases. After maintaining a predetermined lift lower than the lift for a predetermined period, it reaches zero. In the special mode, the exhaust valve is opened during the intake stroke. In the engine described in Patent Document 1, when the internal EGR gas is introduced into the combustion chamber, the controller operates the valve mechanism of the exhaust valve in a special mode. As a result, the exhaust valve is opened during the intake stroke, and part of the burned gas discharged to the exhaust port during the exhaust stroke is reintroduced into the combustion chamber.

  In the engine described in Patent Document 1, by adjusting the internal EGR rate (weight ratio of internal EGR gas to the total weight of gas introduced into the combustion chamber), the temperature state in the combustion chamber is appropriately maintained and mixed. Qi is made to burn stably by self-ignition.

Japanese Patent Laying-Open No. 2015-98802

  By the way, in the engine described in Patent Document 1, the operating region in which the air-fuel mixture in the combustion chamber burns by self-ignition is limited to a relatively low load region. On the other hand, when the operating range in which the air-fuel mixture burns by self-ignition is expanded to the high load side, the fuel efficiency performance of the engine is improved.

  Therefore, unlike the naturally aspirated engine described in Patent Document 1, the engine includes a supercharger that supercharges the intake air introduced into the combustion chamber, and copes with an increase in the amount of fuel in a high load region. It is conceivable to supercharge the intake air so that the new air volume increases. As a result, the operating range in which the air-fuel mixture in the combustion chamber burns by self-ignition can be expanded to the high load side.

  However, as described above, when the supercharger is supercharging the intake air, if the exhaust valve is opened during the intake stroke, both the exhaust valve and the intake valve are opened during the intake stroke. As a result, the intake air is blown from the intake side to the exhaust side of the engine. This makes it difficult to accurately adjust the amount of internal EGR gas introduced into the combustion chamber. That is, the controllability of the internal EGR rate is reduced.

  The means for introducing the internal EGR gas into the combustion chamber is not limited to opening the exhaust valve in the intake stroke, as described in Patent Document 1. For example, by closing both the exhaust valve and the intake valve around the exhaust top dead center, it is possible to confine a part of the burned gas in the combustion chamber.

  Accordingly, it is conceivable to switch the means for introducing the internal EGR gas into the combustion chamber according to the operating state of the engine. However, in this case, it should be avoided that the internal EGR rate greatly fluctuates at the timing of switching means.

  Moreover, even if the fluctuation of the internal EGR rate can be suppressed, good engine performance cannot be obtained unless a stable combustion state can be secured.

  The technology disclosed herein has been made in view of such a point, and the purpose thereof is to change the internal EGR in a compression self-ignition engine when the means for introducing the internal EGR gas into the combustion chamber is switched. This is to suppress the fluctuation of the rate and realize stable combustion.

  Specifically, a control device for a compression self-ignition engine disclosed herein includes an engine configured to burn an air-fuel mixture in a combustion chamber by self-ignition, and at least the exhaust valve among an intake valve and an exhaust valve. A variable valve device configured to introduce internal EGR gas into the combustion chamber by changing a valve opening operation of the engine, and outputting a control signal to at least the variable valve device, A controller configured to operate; and a sensor connected to the controller and configured to detect a parameter related to an operating state of the engine and to output a detection signal to the controller.

  The sensor includes a combustion state detection sensor that detects a combustion state in the combustion chamber. The variable valve operating device includes a first mode in which the internal EGR gas is confined in the combustion chamber by closing the exhaust valve before exhaust top dead center, and the exhaust valve is opened in an intake stroke. Thus, the second mode in which the internal EGR gas is introduced into the combustion chamber is switched, and the first mode and the second mode in the operation region in which the engine is operated by self-ignition combustion. Switching to the mode is executed when the internal EGR rate in the combustion chamber when operating in the first mode is the same as the internal EGR rate when operating in the second mode. .

  Then, during the transition of the switching, the controller determines whether or not the combustion state in the combustion chamber is stable based on the detection signal output from the combustion state detection sensor, and the combustion state is not stable. When it determines with it being stable, the combustion stabilization control which stabilizes the said combustion is performed.

  According to this configuration, the variable valve system closes the exhaust valve before exhaust top dead center in the first mode. As a result, part of the burned gas is confined in the combustion chamber. The variable valve system also opens the exhaust valve in the intake stroke in the second mode. Part of the burned gas discharged to the exhaust port during the exhaust stroke is reintroduced into the combustion chamber when the exhaust valve is opened during the intake stroke.

  The switching between the first mode and the second mode is the same between the internal EGR rate in the combustion chamber when operating in the first mode and the internal EGR rate when operating in the second mode. This is executed at the timing at which the first mode and the second mode are continuously switched without changing the internal EGR rate. Therefore, it is possible to suppress the internal EGR rate from fluctuating greatly when switching between the first mode and the second mode.

  However, in the first mode in which both the intake valve and the exhaust valve are closed around the exhaust top dead center, the high temperature burned gas is confined in the combustion chamber, so the temperature in the combustion chamber is mainly confined in the combustion chamber. It is determined by the amount of burned gas (specifically, the inner wall temperature of the heated combustion chamber). On the other hand, in the second mode in which the internal EGR gas is introduced into the combustion chamber, the temperature in the combustion chamber is mainly determined by the temperature of the internal EGR gas, but the temperature of the internal EGR gas varies depending on the operating state of the engine. To do.

  Therefore, even if the internal EGR rate is kept constant when switching between the first mode and the second mode, there is a possibility that combustion may become unstable due to a temperature difference in the combustion chamber immediately after switching.

  On the other hand, according to this configuration, the combustion state detection sensor detects the combustion state in the combustion chamber, and the controller outputs a detection output from the combustion state detection sensor at the transition between the first mode and the second mode. Based on the signal, for example, for each combustion cycle, it is determined whether the combustion state in the combustion chamber is stable. Thus, for example, when it is determined that the combustion state is unstable in the combustion cycle immediately after switching, combustion stabilization control is performed to stabilize combustion in the subsequent combustion cycle.

  Therefore, even when the combustion state becomes unstable due to the temperature difference in the combustion chamber during the transition between the first mode and the second mode, the combustion can be stabilized immediately, so that smooth It can be switched, and good engine performance can be realized.

  For the combustion stabilization control, for example, internal EGR increase control in which the introduction amount of the internal EGR gas is increased by the variable valve operating device can be used.

  The unstable combustion state is usually a state where combustion is weak and sufficient combustion heat is not obtained. Therefore, combustion can be stabilized by temporarily increasing the introduction amount of the internal EGR gas and increasing the temperature in the combustion chamber during the transition of switching. In the first mode, it is only necessary to increase the amount of burnt gas confined in the combustion chamber, so it is conceivable to advance the valve timing of the exhaust valve by a variable valve operating device. Further, in the second mode, it is only necessary to increase the amount of internal EGR gas introduced. Therefore, it is conceivable that the exhaust valve is largely closed in the intake stroke by the variable valve operating device.

  In addition, fuel injection control can also be used for combustion stabilization control. That is, an injector for injecting fuel into the combustion chamber in a plurality of times from the intake stroke to the compression stroke is further provided, and consists of the ratio of the fuel injection amount on the retard side to the fuel injection amount on the advance side. Combustion stabilization control can also be performed by split injection rate change control that increases the split injection rate.

  That is, since the fuel injection control is excellent in responsiveness, it can be controlled with little time lag. Therefore, the split injection rate change control can perform the combustion stabilization control more quickly and with higher accuracy than the internal EGR increase control. As the fuel is injected on the advance side, that is, on the intake stroke side, the fuel spreads into the combustion chamber. On the other hand, the fuel is injected on the retard side, that is, on the compression stroke side, the fuel is injected into the combustion chamber. It is unevenly distributed and stratified. Thereby, stratified combustion is promoted. If stratified combustion is promoted, combustion can be stabilized even when the fuel amount is the same and the temperature in the combustion chamber is relatively low. Therefore, combustion can also be stabilized by increasing the divided injection rate.

  In particular, the combustion stabilization control includes both the internal EGR increase control and the split injection rate change control, and the controller stabilizes the combustion state in the combustion chamber by performing these controls in cooperation. It is preferable to decrease the split injection rate.

  As the stratified combustion is promoted, the unburned fuel is increased, which is disadvantageous compared with the internal EGR increase control from the viewpoint of improving the fuel efficiency and the exhaust gas performance. Therefore, both the internal EGR increase control and the split injection rate change control are performed in a coordinated manner, and as the combustion stabilizes, the split injection rate is decreased, that is, the stratified combustion is reduced, so that such disadvantage is also improved. it can.

  According to the disclosed control device for a compression self-ignition engine, it is possible to suppress fluctuations in the internal EGR rate and instability of combustion during a transition in which the means for introducing the internal EGR gas into the combustion chamber is switched. Can be realized.

FIG. 1 is a diagram illustrating a configuration of a compression self-ignition engine. FIG. 2 is a block diagram illustrating the configuration of the control device for the compression self-ignition engine. FIG. 3 is a diagram illustrating an engine operating region. FIG. 4 is a diagram exemplifying changes in the fuel injection ratio between the first injection and the second injection, the start timing of the first injection, and the start timing of the second injection with respect to the engine load. FIG. 5 is a diagram illustrating the configuration of the valve operating mechanism of the exhaust valve. FIG. 6 is a diagram illustrating a change in the valve timing of the exhaust valve and a change in the lift amount of the intake valve. FIG. 7 is a diagram illustrating the valve lift relationship between the intake valve and the exhaust valve in each operation region. FIG. 8 is a diagram illustrating changes in internal EGR, exhaust valve timing, and exhaust valve closing timing with respect to engine load. FIG. 9 is a cross-sectional view illustrating changes in the lift amount of the intake valve, the closing timing of the intake valve, the opening timing of the intake valve, and the supercharging pressure with respect to the level of the engine load. FIG. 10 is a diagram illustrating the lift state of the exhaust valve and the lift state of the intake valve in the non-supercharged CI region. FIG. 11 is a diagram illustrating the gas composition in the combustion chamber in the non-supercharged CI region and the supercharged CI region. FIG. 12 is a diagram illustrating another example of changes in the closing timing of the intake valve and the boost pressure with respect to the level of the engine load. FIG. 13 is a diagram illustrating the relationship between the valve timing of the exhaust valve and the internal EGR rate when shifting from the non-supercharged CI region to the supercharged CI region. FIG. 14 is a flowchart regarding a control example of the variable valve system and the supercharging system when the engine operating state shifts between the non-supercharging CI region and the supercharging CI region. FIG. 15 is a flowchart relating to combustion stabilization control. FIG. 16 is a flowchart relating to combustion stabilization control.

  Hereinafter, an embodiment of a control device for a compression self-ignition engine will be described in detail with reference to the drawings. The following description is an example of a control device for a compression self-ignition engine. FIG. 1 is a diagram illustrating a configuration of a compression self-ignition engine. FIG. 2 is a block diagram illustrating the configuration of the control device for the compression self-ignition engine.

  The engine 1 is mounted on a four-wheeled vehicle. The vehicle travels when the engine 1 is driven. The fuel of the engine 1 is gasoline in this configuration example. The fuel may be gasoline containing bioethanol or the like. The fuel of the engine 1 may be any fuel as long as it is a liquid fuel containing at least gasoline.

(Engine configuration)
The engine 1 includes a cylinder block 12 and a cylinder head 13 placed on the cylinder block 12. A plurality of cylinders 11 are formed inside the cylinder block 12. In FIG. 1, only one cylinder 11 is shown. The engine 1 is a multi-cylinder engine.

  A piston 3 is slidably inserted in each cylinder 11. The piston 3 is connected to the crankshaft 15 via a connecting rod 14. The piston 3 forms a combustion chamber 17 together with the cylinder 11 and the cylinder head 13. The “combustion chamber” is not limited to the meaning of the space formed at the timing when the piston 3 reaches the compression top dead center. The term “combustion chamber” may be used in a broad sense. That is, the “combustion chamber” may mean a space formed by the piston 3, the cylinder 11, and the cylinder head 13 regardless of the position of the piston 3. Although detailed illustration is omitted, the upper surface of the piston 3 is a flat surface. A cavity is formed on the upper surface of the piston 3. The lower surface of the cylinder head 13, that is, the ceiling surface of the combustion chamber 17, is constituted by two inclined surfaces. The combustion chamber 17 has a so-called pent roof shape.

  The geometric compression ratio of the engine 1 is set high for the purpose of improving the theoretical thermal efficiency and stabilizing the CI (Compression Ignition) combustion (that is, self-ignition combustion) described later. Specifically, the geometric compression ratio of the engine 1 is 17 or more. The geometric compression ratio may be 18, for example. What is necessary is just to set a geometric compression ratio suitably in the range of 17-20.

  Two intake ports 18 are formed in the cylinder head 13 for each cylinder 11 (note that only one intake port 18 is shown in FIG. 1). The intake port 18 communicates with the combustion chamber 17. An intake valve 21 is disposed in the intake port 18. The intake valve 21 opens and closes between the combustion chamber 17 and the intake port 18. The intake valve mechanism opens and closes the intake valve 21 at a predetermined timing. In this configuration example, the intake valve mechanism has an intake valve VVT (Variable Valve Timing) 231 as a variable valve system and a timing change mechanism, as shown in FIG. The intake VVT 231 is configured to continuously change the rotation phase of the intake camshaft within a predetermined angle range. The opening timing and closing timing of the intake valve 21 change continuously. The intake VVT 231 may be, for example, an electric VVT or a hydraulic VVT.

  In this configuration example, the intake valve mechanism also has an intake CVVL (Continuously Variable Valve Lift) 232 as a variable valve system and a lift changing mechanism, as shown in FIG. The intake CVVL 232 is configured to continuously change the lift amount of the intake valve 21 within a predetermined range. The intake CVVL 232 may adopt a known configuration as appropriate. The configuration of the intake CVVL 232 is arbitrary.

  The cylinder head 13 is also formed with two exhaust ports 19 for each cylinder 11 (note that only one exhaust port 19 is shown in FIG. 1). The exhaust port 19 communicates with the combustion chamber 17. An exhaust valve 22 is disposed in the exhaust port 19. The exhaust valve 22 opens and closes between the combustion chamber 17 and the exhaust port 19. The exhaust valve mechanism opens and closes the exhaust valve 22 at a predetermined timing. In this configuration example, the exhaust valve mechanism includes an exhaust VVT 241 as a variable valve system and a timing changing mechanism, as shown in FIG. The exhaust VVT 241 is configured to continuously change the rotational phase of the exhaust camshaft within a predetermined angle range. The opening timing and closing timing of the exhaust valve 22 change continuously. The exhaust VVT 241 may be an electric VVT or a hydraulic VVT.

  In this configuration example, the exhaust valve mechanism has an exhaust VVL 242 as a variable valve system as shown in FIG. Although the exhaust VVL 242 will be described in detail later, the exhaust valve 22 is configured to change the lift amount of the exhaust valve 22 by opening the exhaust valve 22 by one of the first cam portion 811 and the second cam portion 812. Has been.

  Although details will be described later, the engine 1 adjusts the length of the overlap period related to the opening timing of the intake valve 21 and the closing timing of the exhaust valve 22 by the intake valve mechanism and the exhaust valve mechanism. As a result, the burned gas in the combustion chamber 17 is scavenged, or the hot burned gas is confined in the combustion chamber 17 (that is, the internal EGR (Exhaust Gas Recirculation) gas is introduced into the combustion chamber 17). Or). Further, when the engine 1 is in a predetermined operation state, the exhaust valve 22 is opened in the intake stroke by the exhaust valve mechanism. As a result, the internal EGR gas is introduced into the combustion chamber 17.

  An injector 6 is attached to the cylinder head 13 for each cylinder 11. The injector 6 is configured to inject fuel directly into the combustion chamber 17. The injector 6 is disposed such that its central axis is along the central axis of the cylinder 11. The injector 6 faces the cavity. The central axis of the injector 6 may be shifted from the central axis of the cylinder 11.

  Although not shown in detail, the injector 6 is constituted by a multi-injection type fuel injection valve having a plurality of injection holes. The injector 6 injects the fuel so that the fuel spray spreads radially from the center of the combustion chamber 17.

  The injector 6 is not limited to a multi-hole injector. The injector 6 may employ an external valve opening type injector.

  A fuel supply system 61 is connected to the injector 6. The fuel supply system 61 includes a fuel tank 63 configured to store fuel, and a fuel supply path 62 that connects the fuel tank 63 and the injector 6 to each other. A fuel pump 65 and a common rail 64 are interposed in the fuel supply path 62. The fuel pump 65 pumps fuel to the common rail 64. In this configuration example, the fuel pump 65 is a plunger-type pump driven by the crankshaft 15. The common rail 64 is configured to store the fuel pumped from the fuel pump 65 at a high fuel pressure. When the injector 6 is opened, the fuel stored in the common rail 64 is injected into the combustion chamber 17 from the injection port of the injector 6. The fuel supply system 61 is configured to be able to supply high pressure fuel of 30 MPa or more to the injector 6. The maximum fuel pressure of the fuel supply system 61 may be about 120 MPa, for example. The pressure of the fuel supplied to the injector 6 may be changed according to the operating state of the engine 1. The configuration of the fuel supply system 61 is not limited to the above configuration.

  A spark plug 25 is attached to the cylinder head 13 for each cylinder 11. The spark plug 25 forcibly ignites the air-fuel mixture in the combustion chamber 17. In this configuration example, the spark plug 25 is disposed on the intake side across the central axis of the cylinder 11. The spark plug 25 is adjacent to the injector 6. The spark plug 25 is located between the two intake ports 18. The spark plug 25 is attached to the cylinder head 13 so as to be inclined from the top to the bottom toward the center of the combustion chamber 17. Although not shown in detail, the electrode of the spark plug 25 faces the combustion chamber 17 and is located near the ceiling surface of the combustion chamber 17.

  An intake passage 40 is connected to one side of the engine 1. The intake passage 40 communicates with the intake port 18 of each cylinder 11. The intake passage 40 is a passage through which gas introduced into the combustion chamber 17 flows. An air cleaner 41 that filters fresh air is disposed at the upstream end of the intake passage 40. A surge tank 42 is disposed near the downstream end of the intake passage 40. Although the detailed illustration is omitted, the intake passage 40 downstream of the surge tank 42 constitutes an independent passage branched for each cylinder 11. The downstream end of the independent passage is connected to the intake port 18 of each cylinder 11.

  A throttle valve 43 is disposed between the air cleaner 41 and the surge tank 42 in the intake passage 40. The throttle valve 43 is configured to adjust the amount of fresh air introduced into the combustion chamber 17 by adjusting the opening of the valve.

  A supercharger 44 is also arranged in the intake passage 40 downstream of the throttle valve 43. The supercharger 44 is configured to supercharge intake air introduced into the combustion chamber 17. In this configuration example, the supercharger 44 is a mechanical supercharger that is driven by the engine 1. The mechanical supercharger 44 may be, for example, a roots type. The configuration of the mechanical supercharger 44 may be any configuration. The mechanical supercharger 44 may be a Rishorum type or a centrifugal type.

  An electromagnetic clutch 45 is interposed between the supercharger 44 and the engine 1. The electromagnetic clutch 45 transmits a driving force from the engine 1 to the supercharger 44 between the supercharger 44 and the engine 1 or interrupts the transmission of the driving force. As will be described later, when the ECU 10 switches between disconnection and connection of the electromagnetic clutch 45, the supercharger 44 is switched on and off. That is, in the engine 1, the supercharger 44 can switch between supercharging the gas introduced into the combustion chamber 17 and the supercharger 44 not supercharging the gas introduced into the combustion chamber 17. It is configured to be able to.

  An intercooler 46 is disposed downstream of the supercharger 44 in the intake passage 40. The intercooler 46 is configured to cool the gas compressed in the supercharger 44. The intercooler 46 may be configured to be, for example, a water cooling type.

  A bypass passage 47 is connected to the intake passage 40. The bypass passage 47 connects the upstream portion of the supercharger 44 and the downstream portion of the intercooler 46 in the intake passage 40 so as to bypass the supercharger 44 and the intercooler 46. An air bypass valve 48 is disposed in the bypass passage 47. The air bypass valve 48 adjusts the flow rate of the gas flowing through the bypass passage 47.

  When the supercharger 44 is turned off (that is, when the electromagnetic clutch 45 is disconnected), the air bypass valve 48 is fully opened. As a result, the gas flowing through the intake passage 40 bypasses the supercharger 44 and is introduced into the combustion chamber 17 of the engine 1. The engine 1 is operated in a non-supercharged state, that is, in a natural intake state.

  When the supercharger 44 is turned on (that is, when the electromagnetic clutch 45 is connected), part of the gas that has passed through the supercharger 44 flows back upstream of the supercharger 44 through the bypass passage 47. To do. Since the reverse flow rate can be adjusted by adjusting the opening degree of the air bypass valve 48, the supercharging pressure of the gas introduced into the combustion chamber 17 can be adjusted. In this configuration example, a supercharging system 49 is configured by the supercharger 44, the electromagnetic clutch 45, the bypass passage 47, and the air bypass valve 48.

  An exhaust passage 50 is connected to the other side of the engine 1. The exhaust passage 50 communicates with the exhaust port 19 of each cylinder 11. The exhaust passage 50 is a passage through which exhaust gas discharged from the combustion chamber 17 flows. Although the detailed illustration is omitted, the upstream portion of the exhaust passage 50 constitutes an independent passage branched for each cylinder 11. The upstream end of the independent passage is connected to the exhaust port 19 of each cylinder 11. An exhaust gas purification system having one or more catalytic converters 51 is disposed in the exhaust passage 50. The catalytic converter 51 includes a three-way catalyst. Note that the exhaust gas purification system is not limited to the one containing only the three-way catalyst.

  An EGR passage 52 constituting an external EGR system is connected between the intake passage 40 and the exhaust passage 50. The EGR passage 52 is a passage for returning a part of burned gas to the intake passage 40. The upstream end of the EGR passage 52 is connected to the upstream of the catalytic converter 51 in the exhaust passage 50. The downstream end of the EGR passage 52 is connected to the upstream side of the surge tank 42 in the intake passage 40.

  A water-cooled EGR cooler 53 is disposed in the EGR passage 52. The EGR cooler 53 is configured to cool the burned gas. An EGR valve 54 is also disposed in the EGR passage 52. The EGR valve 54 is configured to adjust the flow rate of burnt gas flowing through the EGR passage 52. By adjusting the opening degree of the EGR valve 54, the recirculation amount of the cooled burned gas, that is, the external EGR gas can be adjusted.

  In this configuration example, the EGR system 55 includes an external EGR system that includes an EGR passage 52 and an EGR valve 54, and an internal EGR system that includes the intake valve mechanism and the exhaust valve mechanism described above. And is composed of.

  As shown in FIG. 2, the control device for the compression self-ignition engine includes an ECU (Engine Control Unit) 10 for operating the engine 1. The ECU 10 is a controller based on a well-known microcomputer, and includes a central processing unit (CPU) 101 that executes a program and, for example, a RAM (Random Access Memory) and a ROM (Read Only Memory). And a memory 102 for storing programs and data, and an input / output bus 103 for inputting / outputting electrical signals. The ECU 10 is an example of a controller.

  As shown in FIGS. 1 and 2, various sensors SW <b> 1 to SW <b> 16 are connected to the ECU 10. Sensors SW1-SW16 output a detection signal to ECU10. The sensors include the following sensors.

  That is, the air flow sensor SW1 that is disposed downstream of the air cleaner 41 in the intake passage 40 and detects the flow rate of fresh air flowing through the intake passage 40, the first intake temperature sensor SW2 that detects the temperature of fresh air, and the intake passage The first pressure sensor SW3 that is disposed upstream of the supercharger 44 at 40 and detects the pressure of the gas flowing into the supercharger 44, is connected to the bypass passage 47 downstream of the supercharger 44 in the intake passage 40 and the bypass passage 47. The second intake air temperature sensor SW4 that is arranged upstream of the position and detects the temperature of the gas flowing out of the supercharger 44, is attached to the surge tank 42, and detects the pressure of the gas downstream of the supercharger 44. Second pressure sensor SW5, a finger pressure sensor SW attached to the cylinder head 13 corresponding to each cylinder 11 and detecting the pressure in each combustion chamber 17 An exhaust temperature sensor SW7 that is disposed in the exhaust passage 50 and detects the temperature of the exhaust gas discharged from the combustion chamber 17, and is disposed upstream of the catalytic converter 51 in the exhaust passage 50 and detects the oxygen concentration in the exhaust gas. A linear O2 sensor SW8, a lambda O2 sensor SW9 that is disposed downstream of the catalytic converter 51 in the exhaust passage 50 and detects the oxygen concentration in the exhaust gas, and a water temperature sensor SW10 that is attached to the engine 1 and detects the temperature of the cooling water. A crank angle sensor SW11 that is attached to the engine 1 and detects the rotation angle of the crankshaft 15; an accelerator opening sensor SW12 that is attached to the accelerator pedal mechanism and detects the accelerator opening corresponding to the operation amount of the accelerator pedal; It is attached to the engine 1 and the intake camshaft An intake cam angle sensor SW13 that detects the turning angle, an exhaust cam angle sensor SW14 that detects the rotation angle of the exhaust camshaft and is attached to the engine 1, and is disposed in the EGR passage 52, and is upstream and downstream of the EGR valve 54. These are an EGR differential pressure sensor SW15 that detects the pressure, and a fuel pressure sensor SW16 that is attached to the common rail 64 of the fuel supply system 61 and detects the pressure of the fuel supplied to the injector 6.

  The ECU 10 determines the operating state of the engine 1 based on these detection signals and calculates the control amount of each device. The ECU 100 sends control signals relating to the calculated control amount to the injector 6, spark plug 25, intake VVT 231, intake CVVL 232, exhaust VVT 241, exhaust VVL 242, fuel supply system 61, throttle valve 43, EGR valve 54, supercharger 44. Output to the electromagnetic clutch 45 and the air bypass valve 48. For example, the ECU 10 adjusts the opening degree of the EGR valve 54 based on the differential pressure across the EGR valve 54 obtained from the detection signal of the EGR differential pressure sensor SW15, so that the amount of external EGR gas introduced into the combustion chamber 17 is adjusted. Adjust. Details of control of the engine 1 by the ECU 10 will be described later.

(Engine operating range)
FIG. 3 illustrates an operation region of the engine 1. The operating region of the engine 1 is roughly divided into four regions with respect to the load level and the rotational speed level. Specifically, the four regions are a low load SI region including a low load region near idle operation and a low load and high rotational speed region, a high load high load SI region including a fully open load, and a low load SI. A region between the region and the high load SI region, which is a non-supercharged CI region having a relatively low load and a low rotation speed, and a supercharged CI region having a relatively high load. The engine 1 performs combustion by compression self-ignition in the non-supercharged CI region and the supercharged CI region, mainly for the purpose of improving fuel consumption and exhaust gas performance. Further, in the low load SI region and the high load SI region, combustion is performed by spark ignition (that is, Spark Ignition: SI). Hereinafter, combustion modes in the low load SI region, the non-supercharged CI region, the supercharged CI region, and the high load SI region will be described in order.

  The combustion mode when the operating state of the engine 1 is in the low load SI region is SI combustion in which the air-fuel mixture is combusted by flame propagation when the spark plug 25 ignites the air-fuel mixture in the combustion chamber 17. In the low load SI region, the air-fuel ratio A / F of the air-fuel mixture (that is, the mass ratio of fresh air and fuel in the combustion chamber 17) is substantially the stoichiometric air-fuel ratio (that is, A / F = 14.7). ). As the three-way catalyst purifies the exhaust gas discharged from the combustion chamber 17, the exhaust gas performance of the engine 1 is improved. The A / F of the air-fuel mixture may be set within the purification window of the three-way catalyst. Therefore, the excess air ratio λ of the air-fuel mixture may be set to 1.0 ± 0.2. The supercharger 44 does not supercharge intake air, and the opening degree of the throttle valve 43 is adjusted as appropriate. The internal EGR gas is not introduced into the combustion chamber 17.

  When the operating state of the engine 1 is in the non-supercharged CI region, the fuel injection amount is larger than that in the low load SI region. Since the combustion temperature becomes high, by introducing the internal EGR gas into the combustion chamber 17, the temperature of the combustion chamber 17 becomes high, and it is possible to perform self-ignition stably. The engine 1 performs CI combustion in the non-supercharged CI region. When the operating state of the engine 1 is in the non-supercharging CI region, the internal EGR gas is introduced into the combustion chamber 17 while the intake air is not supercharged. The A / F of the air-fuel mixture is made leaner than the stoichiometric air-fuel ratio.

  If the load on the engine 1 is increased and the amount of fuel is further increased, the amount of fresh air introduced into the combustion chamber 17 will be insufficient if it is in a natural intake state. Therefore, when the operating state of the engine 1 is in the supercharging CI region, the supercharger 44 supercharges the intake air introduced into the combustion chamber 17 and the engine 1 performs CI combustion. Also, the internal EGR gas is introduced into the combustion chamber 17 even when the operating state of the engine 1 is in the supercharging CI region. The A / F of the air-fuel mixture is made leaner than the stoichiometric air-fuel ratio. The supercharging CI area corresponds to the first area, and the non-supercharging CI area having a lower load than the supercharging CI area corresponds to the second area.

  When the load on the engine 1 is high, the fuel injection amount is large. Therefore, when CI combustion is performed, it becomes difficult to suppress combustion noise. Further, since the temperature in the combustion chamber 17 becomes high, abnormal combustion such as pre-ignition and knocking is likely to occur even when performing CI combustion. Therefore, the combustion mode when the operating state of the engine 1 is in the high load region (that is, the high load SI region) is SI combustion. The A / F of the air-fuel mixture is made approximately the stoichiometric air fuel ratio. Further, when the operating state of the engine 1 is in the high load SI region, the internal EGR gas is not introduced into the combustion chamber 17, but the supercharger 44 performs supercharging. The high load SI area corresponds to a third area having a higher load than the supercharged CI area.

(Fuel injection control)
FIG. 4 shows changes in the fuel injection mode with respect to the load of the engine 1 at a predetermined engine speed. Hereinafter, the fuel injection mode in each of the four regions will be described in order.

(High load SI area)
As described above, the engine 1 performs SI combustion in the high load SI region, but there is a problem that abnormal combustion such as pre-ignition and knocking is likely to occur due to a high geometric compression ratio and the like. .

  Therefore, the engine 1 is configured to avoid abnormal combustion by devising the form of fuel injection in the high load SI region. Specifically, the ECU 10 injects fuel into the combustion chamber 17 at a high fuel pressure of 30 MPa or more and at a timing within a period from the latter stage of the compression stroke to the early stage of the expansion stroke (hereinafter, this period is referred to as a retard period). Thus, a control signal is output to the fuel supply system 61 and the injector 6. Therefore, as shown in the graph 401, the injection ratio between the first injection and the second injection is 10: 0. The injector 6 injects fuel in a lump. Further, as illustrated in the graph 402, the injection start SOI of the first injection is in the latter half of the compression stroke.

  The ECU 10 also outputs a control signal to the spark plug 25 so that the air-fuel mixture is ignited at a timing near the compression top dead center after fuel injection. In the following description, injecting fuel into the combustion chamber 17 at a high fuel pressure and at a timing within the retard period is referred to as high-pressure retarded injection.

  High pressure retarded injection avoids abnormal combustion by shortening the time during which the air-fuel mixture reacts. That is, the reaction time of the air-fuel mixture includes (1) a period during which the injector 6 injects fuel (that is, an injection period), and (2) after the fuel injection is completed, This is a time obtained by adding the period until formation (that is, the mixture formation period) and (3) the period until SI combustion started by ignition ends ((3) combustion period).

  When fuel is injected into the combustion chamber 17 at a high fuel pressure, the injection period and the mixture formation period become shorter. When the injection period and the air-fuel mixture formation period are shortened, the timing for starting fuel injection can be made closer to the ignition timing. The high pressure retarded injection can be performed at a timing within the retard period from the latter stage of the compression stroke to the early stage of the expansion stroke by increasing the pressure of the fuel.

  When fuel is injected into the combustion chamber 17 at a high fuel pressure, the turbulent energy in the combustion chamber 17 increases. When the fuel injection timing is brought close to the compression top dead center, SI combustion can be started with high turbulent energy in the combustion chamber 17. As a result, the combustion period is shortened.

  The high pressure retarded injection can shorten the injection period, the mixture formation period, and the combustion period. Compared with the case where fuel is injected into the combustion chamber 17 during the intake stroke, the high-pressure retarded injection can greatly shorten the time for the air-fuel mixture to react. In the high pressure retarded injection, the time for which the air-fuel mixture reacts is shortened, so that abnormal combustion can be avoided.

  If the fuel pressure is, for example, 30 MPa or more, the injection period, the mixture formation period, and the combustion period can be effectively shortened. The fuel pressure is preferably set as appropriate according to the properties of the fuel. As an example, the upper limit value of the fuel pressure may be 120 MPa.

(Low load SI area)
In the low load SI region, the ECU 10 outputs a control signal to the injector 6 so that the injector 6 injects fuel during the intake stroke. As shown in the graph 401, the injector 6 injects fuel in a lump. The injector 6 may divide and inject fuel. The fuel injected into the combustion chamber 17 during the intake stroke is diffused by the intake air introduced into the combustion chamber 17. The homogeneity of the air / fuel mixture increases. As a result, the fuel consumption of the engine 1 is improved by reducing the unburned loss. Further, exhaust gas performance is improved by avoiding the generation of smoke.

(Non-supercharged CI area)
The non-supercharged CI area is divided into a low load area and a low load area. In the low load region, as in the low load SI region, the injector 6 injects fuel all at once during the intake stroke. On the other hand, in the low load region, the injector 6 performs split injection including the first injection and the second injection. The injection start SOI of the first injection is the first half of the compression stroke as illustrated in the graph 402. The injection start SOI of the second injection is in the latter half of the compression stroke, as illustrated in the graph 403. Here, the first half of the compression stroke is the first half when the period of the compression stroke is divided into two equal parts, and the second half of the compression stroke is the second half when the period of the compression stroke is divided into two equal parts.

  As illustrated in the graph 401, the injection ratio between the first injection and the second injection gradually increases the injection amount of the second injection and gradually decreases the injection amount of the first injection as the load of the engine 1 increases. . The magnitude relationship between the injection amount of the first injection and the injection amount of the second injection is reversed at a predetermined load.

  In the non-supercharged CI region, the ECU 10 outputs a control signal to the injector 6 so as to inject a predetermined amount of fuel at a predetermined timing according to the load of the engine 1.

(Supercharged CI area)
In the supercharging CI region, the injector 6 performs divided injection including the first injection and the second injection regardless of the load of the engine 1. The injection start SOI of the first injection is the first half of the compression stroke as illustrated in the graph 402. The injection start SOI of the second injection is in the latter half of the compression stroke, as illustrated in the graph 403. The ratio between the injection amount of the first injection and the injection amount of the second injection is constant. The injection amount of the second injection is larger than the injection amount of the first injection.

  In the supercharging CI region, the ECU 10 outputs a control signal to the injector 6 so as to perform two fuel injections of the first injection and the second injection during the compression stroke.

(Operation of intake valve and exhaust valve)
Next, changes in the operation of the intake valve 21 and the exhaust valve 22 with respect to the load of the engine 1 will be described.

  First, details of the configuration of the exhaust VVL 242 will be described with reference to FIG. FIG. 5 shows only a specific cylinder 11 in the engine 1. The cylinder head 13 of the engine 1 is provided with two exhaust valves 22 for each cylinder 11 and a return spring 84 for urging the exhaust valve 22 in the closing direction. Further, an exhaust side camshaft 86 that opens the exhaust valve 22 against the urging force of the return spring 84 via the rocker arm 85 is provided at the upper part of the cylinder head 13. The cam shaft 86 is rotatably supported by a bearing portion 80 provided in the cylinder head 13. The cam shaft 86 is rotationally driven via a chain by a crankshaft 15 (not shown in FIG. 5).

  A cylindrical cam element portion 81 is spline-fitted to the cam shaft 86. The cam element portion 81 is coupled to the cam shaft 86 in the rotational direction so as to rotate integrally, and is fitted so as to be slidable in the axial direction. Although not shown, the cam element portions 81 are arranged in a row in the cam shaft direction on the cam shaft 86 so as to correspond to the cylinders 11. The cam element portion 81 includes a first cam portion 811 and a second cam portion 812 that open and close the exhaust valve 22 by sliding contact with the cam follower 851 of the rocker arm 85. The first cam portion 811 corresponds to a first cam, and the second cam portion 812 corresponds to a second cam. Corresponding to two exhaust valves 22 being provided for one cylinder 11, two first cam portions 811 and two second cam portions 812 are provided for each cylinder 11. It has been.

  The first cam portion 811 and the second cam portion 812 are provided adjacent to each other. As will be described later, the first cam portion 811 is a cam portion selected in the low load SI region, the supercharging CI region, and the high load SI region, and the second cam portion 812 is a non-supercharging CI region. It is a cam part selected in.

  The first cam portion 811 and the second cam portion 812 have a base circle 813 in common, and nose portions having different lift amounts are provided on the base circle 813 with a phase difference. The common base circle 813 means that the diameters of the base circles of the base circle 813 of the first cam portion 811 and the second cam portion 812 are the same.

  End cams 82 are respectively provided at both end portions of the cam element portion 81 in the axial direction. The end face cam 82 is formed such that the lift amount in the axial direction gradually increases from the reference surface with respect to the rotation direction, and returns to the reference surface at the lift end position.

  An electromagnetic actuator 83 is disposed above the cam element portion 81 and at a position corresponding to the end face cams 82 on both sides in the cam shaft direction. The electromagnetic actuator 83 has a pin portion 831 that can project toward the cam shaft 86 when energized. In response to the control signal from the ECU 10, in a state where the electromagnetic actuator 83 is not energized, the pin portion 831 is held at the upper inoperative position by a permanent magnet provided inside the actuator. On the other hand, when the electromagnetic actuator 83 is energized, the pin portion 831 protrudes downward against the permanent magnet and advances to the operating position.

  When the pin portion 831 of the electromagnetic actuator 83 advances to the operating position, the tip portion of the pin portion 831 comes into sliding contact with the corresponding end face cam 82, and the cam element portion 81 is slid in the cam shaft direction.

  Specifically, when the electromagnetic actuator 83 on one end side (that is, the right side in FIG. 5) of the cam element portion 81 is operated and the pin portion 831 is brought into sliding contact with the end face cam 82 on one end side of the cam element portion 81, The cam element portion 81 slides to the other end side in the cam shaft direction (that is, the left side in FIG. 5) and shifts so that the second cam portion 812 is in sliding contact with the cam follower 851 of the rocker arm 85. Further, when the electromagnetic actuator 83 on the other end side of the cam element portion 81 is operated so that the pin portion 831 is brought into sliding contact with the end face cam 82 on the other end side of the cam element portion 81, the cam element portion 81 is The first cam portion 811 is shifted so as to be in sliding contact with the cam follower 851 of the rocker arm 85. As described above, the cam portion that opens and closes the exhaust valve 22 is switched among the first cam portion 811 and the second cam portion 812 by the shift of the cam element portion 81 in the axial direction. The end face cam 82 and the electromagnetic actuator 83 constitute a switching mechanism that switches between the first cam portion 811 and the second cam portion 812.

  Next, the opening operation of the intake valve 21 and the exhaust valve 22 will be described with reference to FIG. As shown in FIG. 6, the lift amount of the intake valve 21 continuously changes between the minimum lift and the maximum lift by the intake CVVL 232. Here, the intake CVVL 232 is configured to change the lift amount of the intake valve 21 while maintaining the valve closing timing of the intake valve 21 constant. Therefore, when the lift amount of the intake valve 21 increases, the valve opening timing of the intake valve 21 advances. Although not shown in FIG. 6, the valve timing of the intake valve 21 is changed within a predetermined range by the intake VVT 231.

  Although the cam profile of the first cam portion 811 in the exhaust VVL 242 is not shown in FIG. 6, the lift amount of the exhaust valve 22 is zero with respect to the progress of the crank angle, like the cam profile of the intake valve 21. It has one cam crest that gradually increases to reach the maximum lift amount, and thereafter the lift amount gradually decreases to zero.

  On the other hand, the cam profile of the second cam portion 812 in the exhaust VVL 242 is substantially constant in the lift of the exhaust valve 22 with respect to the progress of the crank angle on the valve closing side in the lift curve, as illustrated in FIG. A lift shelf 222 is maintained. More specifically, the cam profile of the second cam portion 812 has a large lift portion 221, a lift shelf portion 222, and a small lift portion 223. The large lift portion 221 is a portion where the lift amount of the exhaust valve 22 gradually increases from zero to reach the maximum lift amount as the crank angle progresses, and thereafter the lift amount gradually decreases. The large lift portion 221 corresponds to a first cam peak that opens the exhaust valve 22 in the exhaust stroke. The lift shelf 222 is a portion that continues to the large lift portion 221 and reaches the intake stroke while maintaining the lift of the exhaust valve 22 substantially constant. The small lift portion 223 is a portion that is continuous with the lift shelf portion 222 and is a portion in which the lift amount gradually decreases to zero after the lift amount of the exhaust valve 22 is made larger than that of the lift shelf portion 222. The lift shelf 222 and the small lift 223 are continuous with the first cam peak and correspond to a second cam peak that keeps the exhaust valve 22 open from the exhaust stroke to the intake stroke.

  Here, the broken line in FIG. 6 indicates the size of the gap between the top surface of the piston 3 and the ceiling surface of the combustion chamber 17 at the location where the exhaust valve 22 is disposed, and the lift amount of the lift shelf 222 is The maximum lift amount is set as long as the exhaust valve 22 does not interfere with the top surface of the piston 3. Further, by providing the small lift portion 223, the amount of the exhaust valve 22 opened during the intake stroke (that is, the exhaust valve shown in FIG. 6) while shortening the period from opening to closing of the exhaust valve 22 is shortened. The area of the range formed by the lift curve 22 during the intake stroke can be increased. If the amount by which the exhaust valve 22 is opened in the intake stroke is increased, the amount of internal EGR gas introduced into the combustion chamber 17 is increased. Note that the lift amount of the lift shelf 222 may be smaller than the maximum lift amount. Further, the small lift portion 223 may not be provided. In addition, the second cam portion 812 of the exhaust valve 22 maintains a valve opening state through the lift shelf 222 after opening in the exhaust stroke, and instead of a cam profile that closes in the intake stroke, A cam profile that once closes after opening in the exhaust stroke and then opens again in the intake stroke may be adopted.

  The valve timing of the exhaust valve 22 is changed within a predetermined range by the exhaust VVT 241 as indicated by an arrow in FIG.

  FIG. 7 illustrates the relationship between the valve lifts of the intake valve 21 and the exhaust valve 22 in each operation region. First, the graph 701 is an example of the relationship between the valve lifts of the intake valve 21 and the exhaust valve 22 when the engine 1 is operating in the high load SI region. The intake valve 21 is set to a large lift amount by the intake CVVL 232. Further, the opening timing and closing timing of the intake valve 21 are appropriately adjusted by the intake VVT 231 according to the operating state of the engine 1. On the other hand, the exhaust valve 22 is opened by the first cam portion 811 of the exhaust VVL 242. The exhaust valve 22 opens in the exhaust stroke and closes near the exhaust top dead center. The opening timing and closing timing of the exhaust valve 22 are set to appropriate timings by the exhaust VVT 241. When the engine 1 is operating in the high load SI region, both the intake valve 21 and the exhaust valve 22 are opened around the exhaust top dead center. That is, the variable valve system operates in the third mode in the high load SI region. A combination of providing a positive overlap period for the intake valve 21 and the exhaust valve 22 and supercharging by the supercharger 44 facilitates scavenging of the burned gas in the combustion chamber 17. When the engine 1 is operating in the high load SI region, the internal EGR gas introduced into the combustion chamber 17 is substantially zero. This is advantageous in avoiding abnormal combustion in the high load SI region.

  A graph 704 is an example of the relationship between the valve lifts of the intake valve 21 and the exhaust valve 22 when the engine 1 is operating in the low load SI region. The intake valve 21 is set to have a small lift amount by the intake CVVL 232. Further, the opening timing and closing timing of the intake valve 21 are appropriately adjusted by the intake VVT 231 according to the operating state of the engine 1. On the other hand, the exhaust valve 22 is opened by the first cam portion 811 of the exhaust VVL 242. The opening timing and closing timing of the exhaust valve 22 are set to appropriate timings by the exhaust VVT 241. Even when the engine 1 is operating in the low load SI region, the internal EGR gas introduced into the combustion chamber 17 becomes substantially zero.

  A graph 703 is an example of the relationship between the valve lifts of the intake valve 21 and the exhaust valve 22 when the engine 1 is operating in the non-supercharged CI region. The exhaust valve 22 is opened by the second cam portion 812 of the exhaust VVL 242. Accordingly, the closing timing of the exhaust valve 22 is within the intake stroke period. Further, the opening timing and closing timing of the exhaust valve 22 are adjusted by the exhaust VVT 241 in accordance with the load level of the engine 1 as will be described later. As will be described later, the lift amount of the intake valve 21 is set to a predetermined large lift amount by the intake CVVL 232. Further, the closing timing of the intake valve 21 is set to a predetermined timing by the intake VVT 231, while the opening timing of the intake valve is set according to the lift amount. By opening the exhaust valve 22 in the intake stroke, the burned gas discharged to the exhaust port 19 during the exhaust stroke is reintroduced into the combustion chamber 17. A predetermined amount of internal EGR gas is introduced into the combustion chamber 17. That is, in the non-supercharged CI region, the variable valve system operates in the second mode.

  A graph 702 is an example of the relationship between the valve lifts of the intake valve 21 and the exhaust valve 22 when the engine 1 is operating in the supercharging CI region. The exhaust valve 22 is opened by the first cam portion 811 of the exhaust VVL 242. Further, the opening timing and closing timing of the exhaust valve 22 are adjusted by the exhaust VVT 241 in accordance with the load level of the engine 1 as will be described later. The closing timing of the exhaust valve 22 is set before the exhaust top dead center. As will be described later, the lift amount of the intake valve 21 is changed by the intake CVVL 232 in accordance with the load level of the engine 1. Further, the closing timing of the intake valve 21 is set to a predetermined timing by the intake VVT 231, while the opening timing of the intake valve 21 is set according to the lift amount. The exhaust valve 22 and the intake valve 21 are both closed around the exhaust top dead center. By providing a negative overlap period for the exhaust valve 22 and the intake valve 21, a part of the burned gas is confined in the combustion chamber 17. A predetermined amount of internal EGR gas is introduced into the combustion chamber 17. That is, the variable valve system operates in the first mode in the supercharging CI region.

  Next, the control of the variable valve system of the intake valve 21 and the exhaust valve 22 and the control of the supercharging system 49 according to the load level of the engine 1 will be described in detail with reference to FIGS. 8 and 9. In the following, the control of the variable valve system and the supercharging system 49 in each operation region will be described on the assumption that the engine 1 gradually changes from a low load state to a high load state. In addition, when the engine 1 gradually changes from a high load state to a low load high state, the description below is reversed.

  First, a graph 801 in FIG. 8 shows a change in the internal EGR rate with respect to the load of the engine 1 when the rotational speed of the engine 1 is a predetermined rotational speed. In the low load SI region where the load of the engine 1 is low, the amount of internal EGR gas introduced into the combustion chamber 17 is substantially zero as described above. Therefore, the internal EGR rate is also substantially zero. At this time, the exhaust valve 22 is opened by the first cam portion 811. The exhaust VVT 241 sets the valve timing of the exhaust valve 22 to the retard side as shown in the graph 802. As a result, the valve closing timing (EVC) of the exhaust valve 22 is set in the vicinity of the exhaust top dead center as shown in the graph 803.

  The intake CVVL 232 sets the lift amount of the intake valve 21 to a predetermined small lift amount as shown by a graph 901 in FIG. As shown in the graph 902, the intake VVT 231 sets the closing timing (IVC) of the intake valve 21 before the intake bottom dead center. Accordingly, the opening timing (IVO) of the intake valve 21 is set to a predetermined timing as shown in the graph 903.

  In the low load SI region, the supercharging system 49 does not supercharge intake air (see graph 904). Further, the ECU 10 adjusts the throttle valve 43 to an appropriate opening, and adjusts the amount of fresh air introduced into the combustion chamber 17. Therefore, the supercharging pressure (that is, the intake pressure) is lower than the atmospheric pressure.

  In the non-supercharged CI region where the load of the engine 1 is high, a relatively large amount of internal EGR gas is introduced into the combustion chamber 17 as shown in a graph 801. As shown in the graph 803, the exhaust VVL 242 switches between the first cam portion 811 and the second cam portion 812 at the boundary between the low load SI region and the non-supercharged CI region. That is, when the operating state of the engine 1 shifts from the low load SI region to the non-supercharged CI region, the exhaust VVL 242 switches from the first cam portion 811 to the second cam portion 812. As shown in the graph 803, the valve closing timing (EVC) of the exhaust valve 22 is instantaneously retarded from the vicinity of the exhaust top dead center into the intake stroke period (see also the solid line of the graph 703 in FIG. 7). . As a result, a large amount of internal EGR gas can be introduced into the combustion chamber 17. When the variable valve system operates in the second mode, the internal EGR rate can be increased as compared to when the variable valve system operates in the first mode. In the non-supercharged CI region, since the load of the engine 1 is relatively low, the temperature of the internal EGR gas is relatively low, but a large amount of internal EGR gas is generated by operating the variable valve system in the second mode. Since it can introduce into the combustion chamber 17, the temperature in the combustion chamber 17 can be raised to desired temperature. As a result, CI combustion can be stabilized in the non-supercharged CI region.

  As shown in the graph 802, the exhaust VVT 241 gradually advances the valve timing of the exhaust valve 22 as the load of the engine 1 increases in the non-supercharged CI region. As a result, the valve closing timing of the exhaust valve 22 gradually advances toward the exhaust top dead center within the period of the intake stroke, as shown in the graph 803. Since the valve opening amount of the exhaust valve 22 opened during the intake stroke period decreases (see the broken line in the graph 703), the amount of internal EGR gas introduced into the combustion chamber 17 gradually decreases. Therefore, as shown in a graph 801, in the non-supercharged CI region, the internal EGR rate gradually decreases as the load on the engine 1 increases. As the load on the engine 1 increases, the amount of fuel increases and the temperature in the combustion chamber 17 increases. Even if the internal EGR gas is reduced, the air-fuel mixture can be stably burned by self-ignition. Further, in the non-supercharged CI region where the intake air is not supercharged, the amount of fresh air increases as the internal EGR gas is reduced, so that fresh air commensurate with the fuel amount can be secured.

  The intake VVT 231 retards the valve timing of the intake valve 21 in the non-supercharged CI region. As shown in the graph 901, the lift amount of the intake valve 21 is changed to a predetermined large lift amount in the non-supercharged CI region. By doing so, as shown in FIG. 10, the sum of the valve opening amounts of the intake valve 21 and the exhaust valve 22 during the intake stroke period is increased, so that the pump loss of the engine 1 can be reduced. The predetermined large lift amount may be, for example, larger than the minimum lift amount and smaller than the maximum lift amount. Since the intake CVVL 242 changes the lift amount without changing the closing timing of the intake valve 21, the closing timing of the intake valve 21 is constant in the non-supercharged CI region as shown in the graph 902. As shown in the graph 903, the valve opening timing of the intake valve 21 advances as the lift amount increases.

  In the non-supercharging CI region, the supercharging system 49 does not perform supercharging of intake air (see graph 904). The ECU 10 fully opens the throttle valve 43. Thereby, the supercharging pressure becomes atmospheric pressure.

  In the supercharging CI region where the load on the engine 1 is further increased, the internal EGR rate is gradually decreased as the load on the engine 1 is increased, as shown in a graph 801. The internal EGR rate is continuous at the boundary between the non-supercharged CI region and the supercharged CI region. Accordingly, the internal EGR rate continuously decreases as the engine load increases throughout the non-supercharged CI region and the supercharged CI region. As shown in a graph 801, the internal EGR rate decreases linearly as the load on the engine 1 increases from the non-supercharged CI region to the supercharged CI region.

  As described above, in the non-supercharged CI region, the exhaust VVL 242 opens the exhaust valve 22 with the second cam portion 812, and in the supercharged CI region, the exhaust VVL 242 causes the exhaust valve 22 to move to the first cam portion. The valve is opened at 811 (see also graph 803). The ECU 10 causes the exhaust VVL 242 to switch from the second cam portion 812 to the first cam portion 811 at the timing when the operating state of the engine 1 shifts from the non-supercharged CI region to the supercharged CI region. . It is necessary to prevent the internal EGR rate from changing suddenly before and after the cam switching.

  Although detailed illustration of switching between the first cam portion 811 and the second cam portion 812 is omitted, in the cam element portion 81, a step is formed between the first cam portion 811 and the second cam portion 812. It is done at the timing without. Specifically, as indicated by the white arrow in FIG. 7, the lift amount of the exhaust valve 22 is between the first cam portion 811 and the second cam portion 812 on the valve closing side exceeding the maximum lift. It is performed at a predetermined timing that coincides with each other.

  When the exhaust VVL 242 switches from the second cam portion 812 to the first cam portion 811 when the valve timing of the exhaust valve 22 is at a predetermined advance position indicated by a broken line in the graph 703 of FIG. As shown by a solid line in the graph 702, the lift curve of the exhaust valve 22 is instantaneously switched. As a result, the closing timing of the exhaust valve 22 instantaneously changes from the intake stroke period to the exhaust stroke period. The variable valve system closes the exhaust valve 22 before exhaust top dead center from the second mode in which the internal EGR gas has been introduced into the combustion chamber 17 by opening the exhaust valve 22 in the intake stroke. Switching to the first mode in which the internal EGR gas is confined in the combustion chamber 17 by the valve. Although the means for introducing the internal EGR gas into the combustion chamber changes, the state of introducing the internal EGR gas into the combustion chamber can continue.

  At the timing of transition from the non-supercharged CI region to the supercharged CI region, the intake CVVL 232 decreases the lift amount of the intake valve 21 from the large lift amount to a predetermined lift amount as shown in the graph 901. As a result, as shown in a graph 903, the valve opening timing of the intake valve 21 is changed to the retard side. By doing so, as shown in the graph 702 of FIG. 7, the period from the closing timing of the exhaust valve 22 to the exhaust top dead center is equal to the period from the exhaust top dead center to the opening timing of the intake valve 21. . When the negative overlap period of the intake valve 21 and the exhaust valve 22 is provided, the pump loss can be reduced.

  The supercharging system 49 performs supercharging in the supercharging CI region. Specifically, when the ECU 10 connects the electromagnetic clutch 45, the supercharger 44 starts supercharging. As shown in the graph 904, the supercharging pressure gradually increases as the load on the engine 1 increases. The supercharging pressure is adjusted by the ECU 10 by adjusting the opening degree of the air bypass valve 48.

  If the exhaust VVL 242 operates in the second mode in the supercharging CI region, when both the exhaust valve 22 and the intake valve 21 are open in the intake stroke, the combustion chamber 17 The intake air flowing through to the exhaust side through the air will occur. This makes it difficult to accurately adjust the amount of internal EGR gas introduced into the combustion chamber 17. That is, the controllability of the internal EGR rate is reduced.

  On the other hand, by operating the exhaust VVL 242 in the first mode in the supercharging CI region, both the exhaust valve 22 and the intake valve 21 are not opened during the intake stroke. can do. This improves the controllability of the internal EGR rate.

  Further, when the exhaust VVL 242 is operated in the first mode, it becomes difficult to introduce a large amount of internal EGR gas into the combustion chamber 17, but in the supercharging CI region, the load of the engine 1 is relatively high. Since this is the region, the temperature in the combustion chamber 17 required for stabilizing the CI combustion is relatively low. Further, in the supercharging CI region, the amount of fuel is relatively large, so the temperature of the internal EGR gas is relatively high. Therefore, even if the amount of internal EGR gas introduced into the combustion chamber 17 is small, the temperature in the combustion chamber 17 can be increased to a desired temperature.

  In the supercharging CI region, the exhaust VVT 241 gradually retards the valve timing of the exhaust valve 22 as the load on the engine 1 increases as shown in the graph 802. As a result, the valve closing timing of the exhaust valve 22 approaches the exhaust top dead center as shown by the broken line in the graph 803, so the negative overlap period is shortened and the internal EGR rate gradually decreases.

  In the supercharging CI region, as shown in the graph 901, the intake CVVL 232 gradually increases the lift amount of the intake valve 21 as the load on the engine 1 increases. Thus, the closing timing of the intake valve 21 remains constant as shown in the graph 902, and the opening timing of the intake valve 21 gradually advances as shown in the graph 903 (graph 702 in FIG. 7). (See also arrow).

  Thereby, the amount of fresh air introduced into the combustion chamber 17 can be increased so as to correspond to the increase in the fuel amount. FIG. 11 is a comparison between an example of the gas composition in the combustion chamber 17 in the non-supercharged CI region and an example of the gas composition in the combustion chamber 17 in the supercharged CI region. In each of the non-supercharged CI region and the supercharged CI region, the G / F of the air-fuel mixture (that is, the mass ratio of all the gases in the combustion chamber 17 to the fuel) is constant regardless of the load level of the engine 1. On the other hand, the A / F of the air-fuel mixture increases as the load on the engine 1 increases. As described above, in the non-supercharging CI region and the supercharging CI region, the A / F of the air-fuel mixture is leaner than the stoichiometric air-fuel ratio (that is, the air excess ratio λ of the air-fuel mixture exceeds 1). ).

  In the high load SI region, as described above, the internal EGR rate is substantially zero (see graph 801). The exhaust VVT 241 makes the valve timing of the exhaust valve 22 the most retarded as shown in the graph 802, whereby the closing timing of the exhaust valve 22 is retarded from the exhaust top dead center as shown in the graph 803. Side (that is, in the intake stroke). In the high load SI region, the exhaust VVT 241 makes the valve timing of the exhaust valve 22 constant regardless of the load level of the engine 1 (see graphs 802 and 803).

  In the high load SI region, the A / F of the air-fuel mixture becomes substantially the stoichiometric air-fuel ratio. The amount of fresh air introduced into the combustion chamber 17 must be reduced at the timing when the operating state of the engine 1 shifts from the supercharged CI region where the A / F is leaner than the stoichiometric air-fuel ratio to the high load SI region. I must. The intake CVVL 242 temporarily reduces the lift amount of the intake valve 21 as shown in the graph 901. Accordingly, the valve opening timing of the intake valve 21 is once retarded as shown in the graph 903. In the high load SI region, the intake CVVL 242 gradually increases the lift amount of the intake valve 21 as the load on the engine 1 increases. Accordingly, the opening timing of the intake valve 21 gradually advances as the load on the engine 1 increases.

  Further, in the high load SI region, the intake VVT 231 gradually retards the closing timing of the intake valve 21 as the load on the engine 1 increases as shown in the graph 902. The intake valve 21 is delayed and closed in the compression stroke, and the amount of delayed closing of the intake valve 21 gradually increases.

  The supercharging system 49 temporarily reduces the supercharging pressure to adjust the amount of fresh air at the timing when the operating state of the engine 1 shifts from the supercharging CI region to the high load SI region (graph) 904). The supercharging system 49 gradually increases the supercharging pressure as the load on the engine 1 increases in the high load SI region.

  As the load on the engine 1 increases, the amount of slow closing of the intake valve 21 is increased, so that the occurrence of knocking can be avoided. On the other hand, since the supercharging pressure needs to be increased by the amount that the intake valve 21 is closed late, the work of the supercharger 44 is lost.

  Here, as shown in a graph 1201 in FIG. 12 as a modified example, the intake VVT 231 may keep the closing timing of the intake valve 21 constant in the high load SI region. In this case, it is sometimes necessary to retard the ignition timing in order to avoid the occurrence of knocking. On the other hand, as shown in the graph 1202, the supercharging system 49 can relatively reduce the supercharging pressure. The broken line in the graph 1202 corresponds to the solid line in the graph 904 in FIG.

  The closing timing of the intake valve 21 by the intake VVT 231 may be set as appropriate in either the graph 902 of FIG. 9 or the graph 1201 of FIG. 12 according to the efficiency of the supercharger 44. . Further, either the graph 902 or the graph 1201 may be selected according to the rotational speed of the engine 1.

(Cam switching control at the boundary between the non-supercharged CI region and the supercharged CI region)
FIG. 13 shows the relationship between the valve timing of the exhaust valve 22 by the exhaust VVT 241 and the internal EGR rate in the non-supercharged CI region and the supercharged CI region. In FIG. 13, a line indicated by “NVO internal EGR” indicates that the exhaust VVL 242 operates in the first mode, thereby providing a negative overlap period between the intake valve 21 and the exhaust valve 22. This is a relationship between the valve timing of the exhaust valve 22 and the internal EGR rate when the internal EGR gas is introduced. In this case, if the valve timing of the exhaust valve 22 is advanced, the negative overlap period becomes longer, so that the internal EGR rate becomes higher. Therefore, the line indicated as “NVO internal EGR” goes up to the right in FIG.

  When the exhaust VVL 242 operates in the second mode and the exhaust valve 22 is opened in the intake stroke to introduce the internal EGR gas into the combustion chamber 17, the valve timing of the exhaust valve 22 is advanced. Then, since the period during which the exhaust valve 22 is open in the intake stroke is shortened, the internal EGR rate is lowered. On the other hand, if the valve timing of the exhaust valve 22 is retarded, the exhaust valve 22 is opened in the intake stroke. The internal EGR rate increases because the period during which the data is stored increases.

  Here, when the exhaust VVL 242 is operating in the second mode, if the lift amount of the intake valve 21 is changed, the resistance to the gas flow flowing into the combustion chamber 17 through the intake valve 21 changes. The amount of internal EGR introduced into 17 changes. Specifically, when the lift amount of the intake valve 21 is small, the burned gas tends to flow relatively easily through the exhaust valve 22, so that the internal EGR rate increases, and when the lift amount of the intake valve 21 is large, the intake valve 21. Since the fresh air is relatively easy to flow through, the internal EGR rate is lowered. Of the two one-dot chain lines in FIG. 13, the line “maximum lift” indicates the internal EGR rate when the valve timing of the exhaust valve 22 is changed with the lift amount of the intake valve 21 being the maximum lift amount. Showing change. A line indicating “minimum lift” indicates a change in the internal EGR rate when the valve timing of the exhaust valve 22 is changed after setting the lift amount of the intake valve 21 to the minimum lift amount. The lines indicated as “maximum lift” and “minimum lift” are respectively lowered to the right in FIG.

  Further, as indicated by the vertical arrows in FIG. 13, the internal EGR rate can be changed by changing the lift amount of the intake valve 21 even when the valve timing of the exhaust valve 22 is the same.

  As described above, at the boundary between the non-supercharged CI region and the supercharged CI region, the first EGR rate does not change before and after switching between the first cam portion 811 and the second cam portion 812. The cam portion 811 and the second cam portion 812 are switched. Specifically, when the valve timing of the exhaust valve 22 is a predetermined timing by the exhaust VVT 241, the exhaust VVL 242 performs cam switching based on a control signal from the ECU 10. In the non-supercharged CI region, the lift amount of the intake valve 21 is set to a predetermined large lift amount as described above from the viewpoint of reducing pump loss. As the load on the engine 1 increases and the valve timing of the exhaust valve 22 advances, the internal EGR rate gradually decreases as shown by a thick arrow in FIG. The thick arrow indicates the valve timing intersecting with the “NVO internal EGR” line, the internal EGR rate when the exhaust VVL 242 is operating in the second mode, and the exhaust VVL 242 is operating in the first mode. The internal EGR rate at the same time. Accordingly, switching from the second cam portion 812 to the first cam portion 811 is performed at this valve timing. By making the lift amount of the intake valve 21 a large lift amount, as described above, in addition to the effect of reducing pump loss in the non-supercharged CI region, the second cam portion 812 to the first cam portion 811. There is an effect that the amount of advancement of the exhaust valve 22 can be reduced until switching to can be performed.

  For example, in a situation where acceleration is performed as the accelerator pedal is depressed, the operating state of the engine 1 quickly shifts from the non-supercharged CI region to the supercharged CI region, and the supercharging by the supercharging system 49 is promptly performed. To be able to start. It becomes possible to improve the response to the driver's acceleration request.

  Further, when the exhaust VVL 242 is operating in the second mode, the internal EGR rate can be adjusted by adjusting the lift amount of the intake valve 21, so that the valve timing of the exhaust valve 22 is adjusted as described above. In addition, by adjusting the lift amount of the intake valve 21, it is ensured that the internal EGR rate will change when switching between the first cam portion 811 and the second cam portion 812. Can be prevented. Even if the internal EGR rate does not match only by adjusting the valve timing of the exhaust valve 22 due to various factors, the internal EGR rate matches by adjusting the lift amount of the intake valve 21. Thus, switching between the first cam portion 811 and the second cam portion 812 can be performed quickly.

  Next, cam switching control between the non-supercharged CI region and the supercharged CI region will be described with reference to the flowchart shown in FIG. In step S1, the ECU 10 reads the operating state of the engine 1 based on detection signals and the like input from various sensors SW1 to SW16.

  Next, in step S2, ECU10 grasps | ascertains the driving | running state requested | required of the engine 1 based on the detection signal etc. which were input from various sensors SW1-SW16. Further, each device (injector 6, spark plug 25, fuel supply system 61, throttle valve 43 and EGR valve 54) corresponding to the operating state, and each system (variable valve operating system (intake VVT 231, intake CVVL 232, exhaust VVT 241). And a base set of control amounts of the exhaust VVL 242) and the supercharging system 49 (the electromagnetic clutch 45 and the air bypass valve 48).

  Next, in step S3, the ECU 10 determines whether or not the required operation state for the engine 1 is switched between the non-supercharged CI region and the supercharged CI region. When this determination is YES, the process proceeds to step S4, and when it is NO, the process returns to step S1.

  In step S4, the ECU 10 determines whether or not to switch from the non-supercharged CI region to the supercharged CI region. When this determination is YES, the process proceeds to step S5, and when it is NO, the process returns to step S7.

  In step S5, which is switching from the non-supercharged CI region to the supercharged CI region, first, the ECU 10 outputs a control signal so that the exhaust VVL 242 switches from the second cam portion 812 to the first cam portion 811. To do. Thereby, the negative overlap period of the intake valve 21 and the exhaust valve 22 is provided.

  As described above, at the time of switching from the non-supercharging CI region to the supercharging CI region, control is performed so that the internal EGR rate shifts at the same timing so that the temperature state of the combustion chamber 17 does not fluctuate greatly. . However, during the transition, a temperature difference may occur due to the difference in the form of internal EGR gas introduction, and the combustion state may become unstable due to the temperature difference. Therefore, in this engine 1, even in such a case, an auxiliary that enables a stable combustion state to be ensured even during the transition of the switching so that a smooth transition from the non-supercharged CI region to the supercharged CI region can be achieved. Control (combustion stabilization control) is set (details of combustion stabilization control will be described later).

  In subsequent step S6, the ECU 10 outputs a control signal to the supercharging system 49 so as to start supercharging. Thereby, the electromagnetic clutch 45 is connected and the supercharger 44 starts supercharging.

  In this way, after the exhaust VVL 242 switches the cam of the exhaust valve 22, the supercharging system 49 starts supercharging, thereby preventing the intake air from blowing from the intake side to the exhaust side in the supercharging CI region. Can be prevented. As a result, the internal EGR rate can be adjusted with high accuracy.

  In step S7, which is switching from the supercharged CI region to the non-supercharged CI region, the ECU 10 outputs a control signal to the supercharging system 49 so as to stop supercharging. In step S8, the ECU 10 determines whether or not the supercharging pressure has decreased below a predetermined value. While the supercharging pressure has not decreased below the predetermined value, step S8 is repeated. If the supercharging pressure falls below a predetermined value, the process proceeds to step S9.

  In step S9, the ECU 10 outputs a control signal to the exhaust VVL 242 so as to switch from the first cam portion 811 to the second cam portion 812. As a result, the exhaust valve 22 opens in the intake stroke. When the cam is switched, since the supercharging pressure is reduced, it is possible to prevent the intake air from being blown from the intake side to the exhaust side. As a result, the internal EGR rate can be adjusted with high accuracy.

  Even when switching from the supercharged CI region to the non-supercharged CI region, control is performed so that the internal EGR rate is shifted at the same timing. However, even during the transition, the difference in the form of internal EGR gas introduction The combustion state may become unstable due to the influence of the resulting temperature difference. Therefore, the engine 1 is configured to be able to execute the combustion stabilization control even during the transition of switching from the supercharged CI region to the non-supercharged CI region.

<Combustion stabilization control>
As described above, the internal EGR rate in the supercharged CI region where the variable valve system operates in the first mode matches the EGR rate in the non-supercharged CI region where the variable valve system operates in the second mode. Since the operating state of the variable valve system is switched at the timing, the internal EGR rate when transitioning between these regions is continuously switched without fluctuation (see graph 801).

  However, in the first mode in which the exhaust valve 22 is closed before exhaust top dead center, since the high temperature burned gas is trapped in the combustion chamber 17, the temperature of the combustion chamber 17 is the burned temperature trapped in the combustion chamber 17. The amount of gas is determined in detail by the inner wall temperature of the combustion chamber 17 heated by the burned gas. Therefore, it does not fluctuate greatly.

  On the other hand, in the second mode in which the internal EGR gas is introduced into the combustion chamber 17, the temperature of the combustion chamber 17 is mainly determined by the temperature of the internal EGR gas introduced, and the temperature of the internal EGR gas is determined by the engine. 1 depending on the operating state.

  Therefore, even if the internal EGR rate is kept constant during the transition of switching, a difference occurs in the temperature of the combustion chamber 17 between the first mode and the second mode due to the difference in the internal EGR gas introduction mode. obtain. If a difference occurs in the temperature of the combustion chamber 17, the combustion may become unstable.

  Therefore, in this engine 1, in order to improve the unstable combustion state that can occur at the time of transition between the first mode and the second mode, combustion stabilization control can be executed at the time of transition. It is configured. Specifically, immediately after the ECU 10 switches the cams of the exhaust valve 22 (the first cam portion 811 and the second cam portion 812), for example, during a transition between several combustion cycles, in the preceding combustion cycle. Based on the combustion state of the combustion chamber 17, the combustion stabilization control for stabilizing the combustion state of the combustion chamber 17 in the subsequent combustion cycle is executed.

  The switching between the first mode and the second mode may be performed when the non-supercharged CI region is switched to the supercharged CI region or when the supercharged CI region is switched to the non-supercharged CI region (step S5 shown in FIG. 14). And step S9). The former is switching from the second mode to the first mode, and the latter is switching from the first mode to the second mode. After the former switching, burned gas (internal EGR gas) is confined in the combustion chamber 17 by setting the negative overlap period, whereas after the latter switching, the exhaust valve 22 in the intake stroke is changed. The internal EGR gas is introduced into the combustion chamber 17 by opening the valve.

  That is, since the introduction form of the internal EGR gas into the combustion chamber 17 after switching differs depending on the switching direction, the combustion stabilization control is executed according to each case.

  First, FIG. 15 shows the flow of combustion stabilization control in the case of switching from the supercharged CI region, which is the latter, to the non-supercharged CI region. By switching the cam of the exhaust valve 22 shown in step S9 of FIG. 14, the combustion stabilization control shown in FIG. 15 is executed as a subroutine.

  In step S9, when the exhaust VVL 242 is switched from the first cam portion 811 to the second cam portion 812, the ECU 10 detects the detection signal input from the crank angle sensor SW11 and the detection signal input from the finger pressure sensor SW6. Reading is started (step S10). In this embodiment, the crank angle sensor SW11 and the finger pressure sensor SW6 constitute a combustion state detection sensor that detects the combustion state in the combustion chamber 17.

  Then, the ECU 10 grasps the combustion state in the combustion chamber 17 for each combustion cycle based on these detection signals, and whether the combustion state is stable, that is, the combustion chamber 17 has a sufficient temperature and is combusted by self-ignition. Is actually generated stably (step S11).

  As a result, when the ECU 10 determines that the combustion is stable in the combustion cycle immediately after the switching (Yes in step S11), the combustion stabilization control is not performed, and the main routine shown in FIG. Return to control (return to original control based on base set).

  On the other hand, when it is determined that the combustion is not stable (No in step S11), the combustion stabilization control is executed.

  In this embodiment, as combustion stabilization control, control of intake valve 21 and exhaust valve 22 is used to increase the amount of internal EGR gas introduced, and internal EGR increase control and split injection rate change using fuel split injection are performed. Control is used together. By performing these controls in a coordinated manner, appropriate and stable combustion stabilization control can be performed.

  In the non-supercharged CI region after switching, the exhaust valve 22 is opened during the intake stroke and the internal EGR gas is introduced into the combustion chamber 17, so the ECU 10 increases the amount of internal EGR gas introduced. Therefore, control of the exhaust VVT 241 is started as internal EGR increase control so that the closing timing of the exhaust valve 22 is delayed (step S12). As the valve closing timing of the exhaust valve 22 is retarded, the exhaust valve 22 is greatly opened in the intake stroke, so that more internal EGR can be introduced into the combustion chamber 17. As a result, the temperature of the combustion chamber 17 increases and the combustion state can be stabilized.

  At the same time (some time lag may be included), split injection rate change control for increasing the split injection rate is also started (step S13). Since the fuel injection control is more responsive than the control of the exhaust VVT 241, it can be controlled with little time lag. Therefore, according to the divided injection rate change control, combustion stabilization control can be performed more quickly and with higher accuracy than the internal EGR increase control.

  The split injection rate is composed of the ratio of the fuel injection amount on the retard side to the fuel injection amount on the advance side. More specifically, the fuel injection period (the period from the most advanced angle side injection to the most retarded angle side injection) is divided into two equal parts, and the retarded side portion with respect to the fuel injection amount at the advanced angle side portion. It consists of the ratio of fuel injection amount.

  The number of fuel injections may be two or more, and the injection timing may be in any range from the intake stroke to the compression stroke. In the injection pattern based on the base set of the present embodiment, as shown in FIG. 4, the fuel is injected in two parts of the first injection and the second injection in the first half and the second half of the compression stroke. ing. Therefore, according to this, the ratio of the fuel injection amount in the second injection to the fuel injection amount in the first injection corresponds to the split injection rate. The fuel injection in the combustion stabilization control does not necessarily follow the base set. The first injection may be performed in the intake stroke, and is not limited to the first half of the compression stroke.

  As an example, assuming that the injection pattern based on the base set is followed, increasing the divided injection rate means reducing the fuel injection amount in the first injection at the front stage and the fuel injection amount in the second injection at the rear stage accordingly. (The total fuel injection amount is constant). The fuel is injected into the combustion chamber 17 as the fuel is injected on the intake stroke side, whereas the fuel is unevenly distributed in the combustion chamber 17 and is stratified as the fuel is injected on the compression stroke side. Therefore, the combustion mode shifts to stratified combustion as the split injection rate is increased. If stratified combustion is strengthened, combustion can be stabilized even when the fuel amount is the same and the temperature of the combustion chamber 17 is relatively low. Therefore, combustion can also be stabilized by increasing the divided injection rate.

  However, the shift to stratified combustion leads to an increase in unburned fuel. Therefore, from the viewpoint of improving fuel consumption and exhaust gas performance, split injection rate change control is disadvantageous compared to internal EGR increase control. Therefore, the ECU 10 performs both internal EGR increase control and split injection rate change control in a coordinated manner. Then, if the delay of the closing timing of the exhaust valve 22 is advanced by the control of the exhaust VVT 241, the combustion state is stabilized by that amount. The injection rate is gradually reduced (step S14).

  By doing so, an increase in unburned fuel can be suppressed while ensuring the stability of the combustion stabilization control.

  Then, ECU 10 determines whether or not combustion stability is ensured for each combustion cycle (step S15). When the combustion stability is ensured by the ECU 10 (Yes in step S15), the combustion stabilization control ends, and the control returns to the control as the main routine shown in FIG.

  Next, FIG. 16 shows the flow of combustion stabilization control in the case of switching from the non-supercharged CI region, which is the former, to the supercharged CI region. Since the basic control in this case is the same as the control shown in FIG. 15, the same reference numerals are used for the same control, and detailed description thereof is omitted.

  By switching the cam of the exhaust valve 22 in step S5 of FIG. 14, the combustion stabilization control shown in FIG. 16 is executed as a subroutine.

  Specifically, in step S5, when the exhaust VVL 242 is switched from the second cam portion 812 to the first cam portion 811, the ECU 10 receives the detection signal input from the crank angle sensor SW11 and the finger pressure sensor SW6. Reading of the detected signal is started (step S10).

  Then, the ECU 10 grasps the actual combustion state of the combustion chamber 17 for each combustion cycle based on these detection signals, and determines whether or not the combustion state is stable (step S11). As a result, when the ECU 10 determines that the combustion is stable in the combustion cycle immediately after the switching (Yes in step S11), the process returns to the control as the main routine shown in FIG.

  On the other hand, when it is determined that the combustion is not stable (No in step S11), the combustion stabilization control is executed.

  In the supercharged CI region after switching, the burned gas is confined in the combustion chamber 17. Therefore, the ECU 10 starts the control of the exhaust VVT 241 so that the closing timing of the exhaust valve 22 is advanced as the internal EGR increase control in order to increase the introduction amount of the internal EGR gas. Further, the control of the intake VVT 231 is also started so that the valve opening timing of the intake valve 21 is delayed (step S16).

  As the closing timing of the exhaust valve 22 is advanced and the opening timing of the intake valve 21 is retarded, the negative overlap period becomes longer. That is, the amount of burned gas (internal EGR gas amount) confined in the combustion chamber 17 increases. As a result, the temperature of the combustion chamber 17 increases and the combustion state can be stabilized. Although it is preferable to control both the exhaust VVT 241 and the intake VVT 231, only one of them may be controlled.

  At the same time, split injection rate change control for increasing the split injection rate is also started (step S13). The ECU 10 performs both internal EGR increase control and split injection rate change control in a coordinated manner, and controls the exhaust VVT 241 and intake VVT 231 to advance the closing timing of the exhaust valve 22 and delay the opening timing of the intake valve 21. As the angle advances, the combustion state stabilizes accordingly, so that the divided injection rate is gradually lowered in accordance with the progress (step S14).

  Thus, the ECU 10 determines whether or not the combustion stability is ensured for each combustion cycle (step S15). When the combustion stability is secured (Yes in step S15), the combustion stabilization control is performed. Is ended, and the control returns to the main routine shown in FIG.

  Therefore, according to the engine 1, even during the transition between the first mode and the second mode, fluctuations in the internal EGR rate and instability of combustion can be suppressed, so that good engine performance can be realized.

  The technique disclosed here is not limited to being applied to the engine 1 having the above-described configuration. As the configuration of the engine 1, various configurations can be adopted. For example, in the above-described configuration, the engine 1 includes a mechanical supercharger, but instead of this, a turbocharger that is driven by exhaust energy may be included. In the combustion stabilization control, when the temperature of the combustion chamber 17 is too high and the combustion state becomes unstable, the reverse control may be performed.

1 Engine 10 ECU (controller)
17 Combustion chamber 21 Intake valve 22 Exhaust valve 221 Large lift (first cam crest)
222 Lift shelf (second cam mount)
223 Small lift (second cam mountain)
231 Intake VVT (variable valve system, timing change mechanism)
232 Intake CVVL (Variable valve system, lift change mechanism)
241 Exhaust VVT (Variable valve system, timing change mechanism)
242 Exhaust VVL (Variable valve system)
44 Supercharger (mechanical supercharger)
45 Electromagnetic clutch 811 First cam portion (first cam)
812 Second cam portion (second cam)
82 End cam (switching mechanism)
83 Electromagnetic actuator (switching mechanism)
49 Supercharging system SW1 Airflow sensor SW2 First intake temperature sensor SW3 First pressure sensor SW4 Second intake temperature sensor SW5 Second pressure sensor SW6 Finger pressure sensor (combustion state detection sensor)
SW7 Exhaust temperature sensor SW8 Linear O2 sensor SW9 Lambda O2 sensor SW10 Water temperature sensor SW11 Crank angle sensor (combustion state detection sensor)
SW12 Accelerator opening sensor SW13 Intake cam angle sensor SW14 Exhaust cam angle sensor SW15 EGR differential pressure sensor SW16 Fuel pressure sensor

Claims (4)

An engine configured to burn the air-fuel mixture in the combustion chamber by self-ignition;
A supercharging system for supercharging intake air introduced into the combustion chamber;
A variable valve operating device configured to introduce internal EGR gas into the combustion chamber by changing at least an opening operation of the exhaust valve among the intake valve and the exhaust valve;
A controller configured to operate the engine by outputting a control signal to at least the variable valve device;
A sensor connected to the controller and configured to detect a parameter relating to an operating state of the engine and to output a detection signal to the controller;
The sensor includes a combustion state detection sensor for detecting a combustion state in the combustion chamber,
The variable valve operating device includes a first mode in which the internal EGR gas is confined in the combustion chamber by closing the exhaust valve before exhaust top dead center, and the exhaust valve is opened in an intake stroke. By doing so, it is configured to switch between the second mode in which the internal EGR gas is introduced into the combustion chamber,
The engine is have you the operating region for operating the self-ignition combustion, the first mode, the supercharging system is performed in the region for supercharging charge air, and the second mode, the supercharging system Performed in an area where the air supply is not supercharged,
The switching between the first mode and the second mode is made up of an internal EGR rate in the combustion chamber when operating in the first mode and an internal EGR rate when operating in the second mode. Executed when the same,
During the switching transition, the controller determines whether the combustion state in the combustion chamber is stable based on a detection signal output from the combustion state detection sensor, and the combustion state is unstable. A control device for a compression self-ignition engine that performs combustion stabilization control to stabilize the combustion when it is determined that there is. The control device for a compression self-ignition engine according to claim 1,
The combustion stabilization control includes a control unit for a compression self-ignition engine, including an internal EGR increase control for increasing an introduction amount of the internal EGR gas by the variable valve operating device. In the control device for the compression self-ignition engine according to claim 1 or 2,
An injector for injecting fuel into the combustion chamber divided into a plurality of times from the intake stroke to the compression stroke;
The combustion stabilization control includes a split injection rate change control for increasing a split injection rate that is a ratio of a fuel injection amount on the retard side to a fuel injection amount on the advance side. Control device. The control device for a compression self-ignition engine according to claim 1 ,
An injector for injecting fuel into the combustion chamber divided into a plurality of times from the intake stroke to the compression stroke;
The combustion stabilization control is performed both when switching from the first mode to the second mode and when switching from the second mode to the first mode.
Internal EGR increase control for increasing the introduction amount of the internal EGR gas by the variable valve operating device ;
Susumu and split injection rate change control for increasing the split injection ratio consisting of the ratio of the amount of fuel injected at the retard side with respect to the amount of fuel injected at the corner side, both executed,
The controller is a control device for a compression self-ignition engine that reduces the split injection rate as the combustion state in the combustion chamber is stabilized by performing these controls in a coordinated manner.

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