JP2011236835A - Device for control of internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a device for control of an internal combustion engine, which properly controls a temperature of a cylinder, and suppresses a fluctuation of a burning time by a compression ignition, in a compression ignition burning mode.SOLUTION: The device 1 for control of the internal combustion engine 3, in a compression ignition burning mode, sucks again exhaust gas as high temperature gas into a cylinder C, the exhaust gas which is discharged into an exhaust passage 5, by opening and closing an exhaust valve 13 from the end of a suction stroke through to a compression stroke. Then, low temperature gas having a lower temperature than high temperature gas is supplied into the cylinder C so that the high temperature gas is cooled by low temperature gas supplying mechanisms 80 and 90. The amount GAIR2 of low temperature gas supplied from the low temperature gas supplying mechanisms is controlled onto a reduction side, as a burning time expressed by an obtained burning time parameter TEX is earlier.

Description

  The present invention relates to a control device for an internal combustion engine that controls the internal combustion engine in a compression ignition combustion mode in which an air-fuel mixture is combusted by compression ignition.

  As a conventional control device for this type of internal combustion engine, for example, one disclosed in Patent Document 1 is known. First and second exhaust ports are formed in the cylinder of the internal combustion engine. The control device includes a supply passage connected to the intake passage and the first exhaust port, and an on-off valve that opens and closes the supply passage. By opening the on-off valve, fresh air passes through the supply passage. The low temperature gas is supplied to the first exhaust port and filled.

  In this control device, in the compression ignition combustion mode, the first exhaust port is filled with the low temperature gas by opening the on-off valve at a constant timing during the intake stroke. In addition, by opening the exhaust valve at the end of the intake stroke, the exhaust gas discharged into the exhaust passage is re-intaked into the cylinder through the second exhaust port as a high-temperature gas and filled into the first exhaust port. The generated low temperature gas is sucked into the cylinder, and the low temperature gas and the high temperature gas are stratified. The stratified low temperature gas and high temperature gas are combusted by compression ignition in the subsequent compression stroke. This combustion is started by auto-ignition from the hot gas portion and reaches the cold gas portion. As described above, in the compression ignition combustion mode, the low temperature gas is sucked through the first exhaust port so that the combustion by the self ignition is performed slowly, thereby preventing the occurrence of knocking.

International Publication No. 2008/140036 Pamphlet

  However, in the conventional control device, since the on-off valve is opened at a constant timing, the temperature in the cylinder may not be appropriately controlled. For example, when the low temperature gas sucked into the cylinder is excessive, the temperature in the cylinder may be excessively lowered due to the shortage of the high temperature gas, and in this case, the combustion timing is delayed. On the contrary, if the low temperature gas sucked into the cylinder is insufficient, the temperature inside the cylinder may rise too much due to the reinhalation of more high temperature gas. The time will be early. In such a case, the combustion timing tends to vary between combustion cycles, and there is room for improvement in this respect.

  The present invention has been made to solve such a problem, and in the compression ignition combustion mode, the internal combustion engine can appropriately control the temperature in the cylinder and suppress the variation in the combustion timing due to the compression ignition. An object of the present invention is to provide a control device.

  In order to achieve the above object, the invention according to claim 1 has a compression ignition combustion mode in which the air-fuel mixture is combusted by compression ignition as the combustion mode, and controls the internal combustion engine 3 in the compression ignition combustion mode. The control device 1 includes an intake side valve mechanism 40 that draws fresh air into the cylinder C by opening and closing the intake valve 12 during the intake stroke, and an exhaust valve 13 from the end of the intake stroke to the compression stroke. An exhaust gas intake mechanism (EGR intake mechanism 80 in the embodiment (hereinafter, the same in this section) in the embodiment) for inhaling exhaust gas discharged into the exhaust passage 5 again into the cylinder C as a high-temperature gas by opening and closing; A low temperature gas supply mechanism (EGR suction mechanism 80, fresh air supply mechanism) for supplying a low temperature gas having a temperature lower than that of the high temperature gas into the cylinder C in order to cool the high temperature gas. 0) and combustion timing parameter acquisition means (# 1 to # 4 exhaust temperature sensors 26 to 29) for acquiring combustion timing parameters (exhaust temperature TEX) indicating the combustion timing of fresh air and high-temperature gas sucked into the cylinder C As the combustion timing represented by the acquired combustion timing parameter is earlier, the control means (ECU2, FIG. 9) controls the amount of low-temperature gas supplied by the low-temperature gas supply mechanism (secondary fresh air amount GAIR2) to the decreasing side. Step 46, FIG. 10).

  According to the control device for an internal combustion engine, the internal combustion engine is controlled as follows in the compression ignition combustion mode. First, fresh air is sucked into the cylinder by opening the intake valve by the intake side valve mechanism during the intake stroke. Further, from the end of the intake stroke to the compression stroke, the exhaust valve is opened by the exhaust gas intake mechanism, whereby the exhaust gas discharged into the exhaust passage is again sucked into the cylinder as a high temperature gas. These fresh air and hot gas are combusted by compression ignition in the subsequent compression stroke. This combustion starts from the hot gas part and reaches the fresh air part. The high temperature gas is cooled by the low temperature gas supplied into the cylinder by the low temperature gas supply mechanism.

  The combustion timing due to compression ignition changes depending on the temperature in the cylinder, and the temperature in the cylinder mainly changes according to the amount of fuel contained in the hot gas and the temperature of the hot gas. Based on such a viewpoint, according to the present invention, the amount of low-temperature gas is controlled according to the acquired combustion timing parameter representing the combustion timing, so the temperature in the cylinder in the subsequent compression stroke is set to the actual combustion timing. Accordingly, the combustion timing in the compression ignition combustion mode can be appropriately controlled. Thereby, the dispersion | variation in the combustion time between combustion cycles can be suppressed. In addition, when the internal combustion engine has a plurality of cylinders, it is possible to suppress variations in combustion timing among the cylinders by controlling the amount of low-temperature gas for each cylinder.

  In addition, the earlier the combustion timing, the earlier the combustion end timing. Therefore, the cooling loss of the combustion gas in the cylinder in the expansion stroke becomes larger, so that the temperature of the exhaust gas discharged in the subsequent exhaust stroke, and thus the hot gas The temperature of becomes lower. In addition, if the temperature of the hot gas is low, the combustion timing in the next compression stroke is delayed, and accordingly, the cooling loss of the combustion gas in the expansion stroke is reduced, and the temperature of the hot gas is increased, so that the next compression is performed. Combustion time in the stroke becomes earlier. As the combustion timing is advanced, the combustion timing in the next compression stroke is further delayed. As described above, when the combustion timing is early or late, the phenomenon that the combustion timing is advanced or delayed not only in the combustion cycle but also in the subsequent combustion cycles due to the influence thereof is repeated.

  Based on the above viewpoint, according to the present invention, the earlier the combustion timing represented by the acquired combustion timing parameter, the lower the amount of low temperature gas is controlled. Thereby, combustion temperature can be raised by reducing the cooling degree by low temperature gas. As described above, the earlier the combustion timing, that is, the greater the cooling loss, the higher the temperature of the exhaust gas and thus the temperature of the high-temperature gas can be compensated. Can be avoided, and therefore variations in combustion timing can be suppressed.

  According to a second aspect of the present invention, in the control device 1 for the internal combustion engine 3 according to the first aspect, the internal combustion engine 3 further has a spark ignition combustion mode in which the air-fuel mixture is combusted by spark ignition as the combustion mode. The means is characterized in that the amount of low temperature gas is controlled when the combustion mode is switched from the spark ignition combustion mode to the compression ignition combustion mode.

  Combustion by compression ignition is normally performed when the internal combustion engine is in a low-load operation state, and combustion by spark ignition is performed at other times. For this reason, the wall surface temperature of the cylinder is higher when the combustion mode is the spark ignition combustion mode than when the combustion mode is the compression ignition combustion mode. For this reason, during switching from the spark ignition combustion mode of the combustion mode to the compression ignition combustion mode, the temperature in the cylinder tends to increase, and the air-fuel mixture may burn early due to self-ignition. Based on this point of view, according to the present invention, the low temperature gas amount is controlled when the spark ignition combustion mode is switched to the compression ignition combustion mode. It can be avoided. As a result, variation in combustion timing can be suppressed even during switching to the compression ignition combustion mode.

1 is a diagram schematically showing a configuration of an internal combustion engine to which the present invention is applied. It is a block diagram which shows schematic structure of a control apparatus. It is the elements on larger scale of FIG. It is a schematic diagram which shows schematic structure of a switching mechanism. It is a partial expansion perspective view of a piston. It is a main flow which shows a combustion control process. It is an example of the map used by the process of FIG. It is a subroutine which shows CI combustion control processing. It is a main flow which shows a secondary fresh air control process. It is an example of the table used by the process of FIG. It is a subroutine which shows the calculation process of secondary fresh air injection timing. It is an example of the table used by the process of FIG. It is a figure which shows the valve lift curve of the intake valve obtained by the intake side valve mechanism, and the valve lift curve of the exhaust valve obtained by the EGR intake mechanism. It is a figure which shows an example of operation | movement of an intake valve, an exhaust valve, etc. in the switching to CI combustion mode or CI combustion mode. It is a timing chart which shows the example of operation obtained by secondary fresh air control processing.

  Hereinafter, preferred embodiments of the present invention will be described with reference to the drawings. As shown in FIG. 1, an internal combustion engine (hereinafter referred to as “engine”) 3 to which the control device 1 of the present embodiment is applied is a four-cylinder gasoline having # 1 to # 4 (first to fourth) cylinders C. An engine, which is mounted on a vehicle (not shown).

  In this engine 3, the phase of the combustion cycle is shifted by 180 ° in the order of # 1 cylinder C → # 3 cylinder C → # 4 cylinder C → # 2 cylinder C → # 1 cylinder C. Are performed in this order.

  An intake passage 4 and an exhaust passage 5 are connected to the cylinder head 3a of the engine 3, and as shown in FIG. 3, an in-cylinder fuel injection valve 10 and a spark plug 11 are connected to the combustion chamber 3b for each cylinder C. It is attached to face. The in-cylinder fuel injection valve 10 is provided on the lower side of the intake port 4a, and is a direct injection type that directly injects fuel into the combustion chamber 3b. The valve opening time and the valve opening timing of the in-cylinder fuel injection valve 10 are controlled by the ECU 2 described later, whereby the fuel injection amount and the fuel injection timing by the in-cylinder fuel injection valve 10 are controlled. The ignition timing of the spark plug 11 is also controlled by the ECU 2.

  The engine 3 is provided with a port fuel injection valve 9 for each cylinder C in addition to the in-cylinder fuel injection valve 10. Each port fuel injection valve 9 is attached to each branch passage 4b of the intake manifold and faces the intake port 4a. The valve opening time and the valve opening timing of the port fuel injection valve 9 are also controlled by the ECU 2, whereby the fuel injection amount and the fuel injection timing by the port fuel injection valve 9 are controlled.

  In this engine 3, as a combustion mode, a homogeneous mixture generated by injecting fuel from the port fuel injection valve 9 and the in-cylinder fuel injection valve 10 during the intake stroke is burned by spark ignition by the spark plug 11. A spark ignition combustion mode (hereinafter referred to as “SI combustion mode”) and a compression ignition combustion mode (hereinafter referred to as “CI combustion mode”) in which a stratified mixture generated as described later is combusted by self-ignition. The switching is controlled by the ECU 2.

  As shown in FIG. 1, the exhaust passage 5 includes # 1 to # 4 first exhaust passages 5a to 5d connected to # 1 to # 4 cylinders C, and # 1 and # 4 first exhaust passages 5a and 5d, respectively. And the second exhaust passages 6a and 6b connected to the joining portions of the # 2 and # 3 first exhaust passages 5b and 5c, respectively, and the second exhaust passages 6a and 6b connected to the joining portions of the second exhaust passages 6a and 6b. 3 exhaust passages 7 are provided.

  A filter 14 is provided in each of the second exhaust passages 6 a and 6 b and the third exhaust passage 7. The filter 14 reduces the amount of PM discharged into the atmosphere by collecting PM such as soot in the exhaust gas.

  As shown in FIG. 3, the cylinder C is provided with a pair of intake valves 12, 12 and a pair of exhaust valves 13, 13 (only one is shown). The intake valve 12 is opened and closed by an intake side valve mechanism 40, and the exhaust valve 13 is opened and closed by an exhaust side valve mechanism 60. The configurations of the intake side valve mechanism 40 and the exhaust side valve mechanism 60 are the same as those already proposed in Japanese Patent Application No. 2009-168228 by the applicant of the present application. .

  The intake side valve mechanism 40 is a variable mechanism that changes the lift and valve timing of the intake valve 12. Note that the lift of the intake valve 12 and the lift of the exhaust valve 13 to be described later represent the maximum lifts of the intake valve 12 and the exhaust valve 13, respectively.

  The intake side valve mechanism 40 includes a rotatable intake camshaft 41, a pair of intake cams 42 and 42 (only one shown) provided integrally with the intake camshaft 41, an intake control shaft 43, and the intake control shaft 43. Actuator 44 (see FIG. 2), a pair of swing cams 45 and 45 (only one is shown) supported swingably on a support shaft 47, a control arm mechanism 46, an intake cam phase variable mechanism 50, etc. It has. The intake cam phase varying mechanism 50 changes the relative phase of the intake camshaft 41 with respect to a crankshaft (not shown) in a stepless manner.

  The intake camshaft 41 is connected to the crankshaft via an intake sprocket and a timing chain (both not shown), and rotates once every two rotations of the crankshaft.

  The control arm mechanism 46 includes a control arm 46a, a roller 46b, an intake rocker arm 46c, and the like. The control arm 46a is rotatably supported on the eccentric shaft 43a of the intake control shaft 43 at the base end portion. A roller 46b is provided at the other end of the control arm 46a. The control arm 46a is in contact with the swing cam 45 via the roller 46b.

  The intake rocker arm 46c includes a main body portion 46d rotatably supported by the intake control shaft 43, an extension portion 46e extending from the main body portion 46d, and the like. In the extension portion 46e, the roller 46b and the intake valve 12 are provided. It is in contact.

  With the above configuration, when the swing cam 45 is not pressed by the intake cam 42, the intake valve 12 is held in the closed position shown in FIG. When the swing cam 45 is pressed by the intake cam 42 as the intake cam shaft 41 rotates, the swing cam 45 rotates about the support shaft 47 counterclockwise in FIG. At that time, when the roller 46b is pressed by the swing cam 45, the intake rocker arm 46c rotates about the intake control shaft 43 through the roller 46b in the counterclockwise direction in FIG. By pushing down, the intake valve 12 is opened.

  When the intake control shaft 43 is rotated via the actuator 44 described above, the control arm 46 a moves in the left-right direction in FIG. 3 about the eccentric shaft 43. Along with this movement, the contact position of the control arm 46a with the swing cam 45 changes, whereby the lift and valve timing of the intake valve 12 are changed steplessly. More specifically, as shown in FIG. 13, the valve closing timing is earlier as the lift of the intake valve 12 is smaller.

  The exhaust side valve mechanism 60 described above includes a rotatable exhaust cam shaft 61, a pair of exhaust cams 62 and 62 (only one shown) provided integrally with the exhaust cam shaft 61, an exhaust control shaft 63, and the exhaust control. A pair of rocker arms 64 and 64 (only one is shown) that are rotatably supported by the shaft 63 and abut against the upper ends of the exhaust valves 13 and 13, respectively, a roller 65 provided on the rocker arm 64, and an exhaust cam phase variable. A mechanism 70 and the like are provided.

  The exhaust camshaft 61 is connected to the crankshaft via an exhaust sprocket and a timing chain (both not shown), and rotates once every two rotations of the crankshaft. When the exhaust camshaft 61 rotates, the rocker arms 64, 64 are pressed by the exhaust cam 62, and rotate about the exhaust control shaft 63 in the clockwise direction in FIG. 3, thereby opening the exhaust valves 13, 13. Further, the exhaust cam phase variable mechanism 70 changes the opening / closing valve timing of the exhaust valves 13 and 13 steplessly by changing the relative phase of the exhaust camshaft 61 with respect to the crankshaft steplessly.

  Further, the engine 3 includes an EGR intake mechanism 80 for causing the exhaust gas to be again taken into the cylinder C in the CI combustion mode.

  The EGR intake mechanism 80 opens the exhaust valve 13 during a predetermined period during the intake stroke, and the exhaust gas once discharged into the # 1 to # 4 first exhaust passages 5a to 5d of the exhaust passage 5 is discharged into the cylinder. Inhaled into C. As shown in FIGS. 3 and 4, the EGR suction mechanism 80 is provided between the two intake cams 42, 42 and is rotatably supported by the EGR cam 81 integrated with the intake camshaft 41 and the support shaft 47. A rocker cam 82, a control arm 83, a lever 84, a switching mechanism 85, and the like are provided. The EGR cam 81 is in contact with the roller 82 a of the rocker cam 82.

  The control arm 83 is rotatably supported by the eccentric shaft 63a of the exhaust control shaft 63 at the base end portion. A roller 83 a is provided at the other end of the control arm 83. The control arm 83 is in contact with the rocker cam 82 via the roller 83a.

  The lever 84 is provided between the two rocker arms 64 and 64. The lever 84 has a triangular shape, and is pivotally supported by the exhaust control shaft 63 at the apex corner, and abuts against the pressing portion 83b of the control arm 83 at one bottom corner. Yes.

  With the above configuration, when the roller 82a is pressed by the EGR cam 81 as the intake camshaft 41 rotates, the rocker cam 82 rotates about the support shaft 47 in the clockwise direction in FIG. At this time, the roller 83a of the control arm 83 is pressed by the rocker cam 82, whereby the control arm 83 rotates about the exhaust control shaft 63 in the clockwise direction in FIG. 84 is pressed. Thereby, the lever 84 rotates around the exhaust control shaft 63 in the clockwise direction of FIG.

  As shown in FIG. 4, the switching mechanism 85 includes cylinders 86 a to 86 c formed on the rocker arms 64 and 64 and the lever 84, connection pistons 87 and 88 housed in the cylinders 86 a to 86 c, and the like. The cylinder 86a is provided with a return spring 89 that biases the connecting pistons 87 and 88 toward the opposite cylinder 86c. Further, hydraulic pressure is supplied to the cylinder 86 c via an oil passage (not shown) formed in the exhaust control shaft 63. The hydraulic pressure is supplied to the cylinder 86c by the ECU 2 controlling supply / stop of the hydraulic pressure from a pump (not shown) to the oil passage.

  With the above configuration, when no hydraulic pressure is supplied to the cylinder 86c, the connecting piston 87 is accommodated in the cylinder 86b and the connecting piston 88 is accommodated in the cylinder 86c by the urging force of the return spring 89 (FIG. 4A). )). As a result, the rocker arm 64 and the lever 84 are disconnected from each other and become free, so that the movement of the lever 84 is not transmitted to the rocker arm 64 and only the lever 84 is driven by the EGR cam 81.

  On the other hand, when the hydraulic pressure is supplied to the cylinder 86c, the connecting pistons 87 and 88 move toward the cylinder 86a against the urging force of the return spring 89 by this hydraulic pressure, so that the connecting piston 87 is moved to the cylinder 86a and the cylinder 86b. The connecting piston 88 is engaged in a state of straddling the cylinder 86b and the cylinder 86c (FIG. 4B). As a result, the lever 84 and the rocker arm 64 are connected, and the movement of the EGR cam 81 is transmitted to the rocker arm 64 via the lever 84, whereby the exhaust valve 13 is opened and closed with a constant lift and a constant valve timing.

  A plurality of guide walls 3d are attached to the cylinder head 3a. As shown in FIG. 3, the guide wall 3d extends over each of the openings of the # 1 to # 4 exhaust ports 8a to 8d over a part of the circumferential direction and protrudes into the combustion chamber 3b. Each guide wall 3d is disposed on the intake port 4a side, and its length is smaller than the radius of the openings of # 1 to # 4 exhaust ports 8a to 8d. The protruding height of the guide wall 3d is substantially the same as the lift of the exhaust valve 13, and is set to 2 to 3 mm, for example.

  As shown in FIG. 5, a convex portion 3e is formed on the top surface of the piston 3c. The convex portion 3e is arranged so that the center thereof is on the intake valve 12 side. Further, the edge 3g on the intake valve 12 side of the convex portion 3e is formed in a substantially straight line, whereas the edge 3h on the exhaust valve 13 side is curved in a state of being recessed toward the intake valve 12 side. Yes.

  With the above configuration, when the exhaust valve 13 is opened by the EGR intake mechanism 80 in the CI combustion mode, the exhaust gas discharged to the exhaust passage 5 is re-intaken into the cylinder C via the exhaust valve 13. At this time, the exhaust gas is sucked into the cylinder C while being guided downward along the inner wall on the exhaust valve 13 side by the guide wall 3d, and the intake air of the exhaust gas sucked by the convex portion 3e of the piston 3c. Outflow to the valve 12 side is prevented. As a result, the exhaust gas flows into the cylinder C as indicated by the solid line A in FIG. As a result, a high-temperature gas layer T1 of higher-temperature exhaust gas is formed on the exhaust valve 13 side in the cylinder C, and a mixed gas layer T2 of lower-temperature fresh air is formed on the intake valve 12 side. New air and exhaust gas are stratified.

  Further, at the same time when the exhaust valve 13 is opened by the EGR intake mechanism 80, the first exhaust # 1 to # 4 is exhausted from the cylinder C in the exhaust stroke in which the phase of the combustion cycle is shifted by 360 ° with respect to the cylinder C. Pressure is introduced through the passages 5a to 5d. For example, because the phases of the # 1 cylinder C and the # 4 cylinder C are shifted from each other by 360 °, when the # 1 cylinder C is between the intake stroke and the compression stroke, the # 4 cylinder C is in the expansion stroke to the exhaust stroke. Between. For this reason, when the exhaust gas is re-inhaled into the # 1 cylinder C, the exhaust gas is properly re-inhaled into the # 1 cylinder C by the pressure of the exhaust gas discharged from the # 4 cylinder C.

  As shown in FIG. 1, the engine 3 further includes a fresh air supply mechanism 90 for cooling the exhaust gas re-inhaled into the cylinder C. The fresh air supply mechanism 90 includes a supply passage 91 for supplying fresh air in the intake passage 4 to the # 1 to # 4 exhaust ports 8a to 8d, and a supercharger provided in the supply passage 91 in order from the upstream side. 92, a cooler 93, # 1 to # 4 control valves 94a to 94d, and the like. The supply passage 91 branches from the intake passage 4 and is connected to the # 1 to # 4 exhaust ports 8a to 8d on the downstream side. More specifically, the downstream end of the supply passage 91 is branched into # 1 to # 4 branch passages 91a to 91d, and these # 1 to # 4 branch passages 91a to 91d are respectively # 1 to # 4. It faces one of the four exhaust ports 8a to 8d.

  The supercharger 92 supercharges fresh air flowing through the supply passage 91 (hereinafter referred to as “secondary fresh air”). The supercharging pressure of the secondary fresh air by this supercharging is controlled by the ECU 2 so as to be higher than the pressure in the # 1 to # 4 exhaust ports 8a to 8d. The cooler 93 cools the supercharged secondary fresh air.

  The # 1 to # 4 control valves 94a to 94d are provided in the # 1 to # 4 branch passages 91a to 91d, respectively. The valve opening times and valve opening timings of the # 1 to # 4 control valves 94a to 94d are also controlled by the ECU 2, whereby the amount of secondary fresh air supplied to the # 1 to # 4 exhaust ports 8a to 8d ( GAIR2 (hereinafter referred to as “secondary fresh air amount”) and secondary fresh air injection timing (hereinafter referred to as “secondary fresh air injection timing”) TAIR2 are controlled.

  The engine 3 is provided with a crank angle sensor 21 and a cylinder discrimination sensor 30. The crank angle sensor 21 outputs a CKR signal and a TDC signal, which are pulse signals, to the ECU 2 as the crankshaft rotates.

  The CRK signal is output every predetermined crank angle (for example, 1 °). The ECU 2 calculates the engine speed (hereinafter referred to as “engine speed”) NE of the engine 3 based on the CRK signal. The TDC signal is a signal indicating that the piston 3c is in a predetermined crank angle position slightly before the top dead center at the start of the intake stroke in any cylinder C, as in the present embodiment. When the engine 3 has four cylinders, it is output every crank angle of 180 °. The cylinder discrimination sensor 30 outputs a cylinder discrimination signal, which is a pulse signal for discriminating the cylinder C, to the ECU 2.

  An intake air temperature sensor 22 and an air flow sensor 23 are provided in the intake passage 4 in order from the upstream side. The intake air temperature sensor 22 detects a temperature TA (hereinafter referred to as “intake air temperature”) TA in the intake passage 4 and outputs a detection signal to the ECU 2. The air flow sensor 23 detects a fresh air amount GAIR sucked into the engine 3 and outputs a detection signal to the ECU 2.

  The # 1 to # 4 first exhaust passages 5a to 5d of the exhaust passage 5 are provided with # 1 to # 4 exhaust temperature sensors 26 to 29, respectively. These # 1 to # 4 exhaust temperature sensors 26 to 29 respectively control the temperatures in the # 1 to # 4 first exhaust passages 5a to 5d (hereinafter referred to as "# 1 to # 4 exhaust temperatures") TEX1 to TEX4, respectively. The detected signals are output to the ECU 2.

  Further, the ECU 2 receives a detection signal indicating a temperature TW (hereinafter referred to as “engine water temperature”) TW circulating from the water temperature sensor 24 in the cylinder block 3 f of the engine 3 from the accelerator opening sensor 25. Detection signals representing the amount of depression (hereinafter referred to as “accelerator opening”) AP of an accelerator pedal (not shown) are output.

  The ECU 2 is composed of a microcomputer including a CPU, a RAM, a ROM, an I / O interface (all not shown), and the like. The ECU 2 determines the operating state of the engine 3 according to the detection signals of the various sensors 21 to 30 described above, and sets the combustion mode of the engine 3 to the SI combustion mode or the CI combustion according to the determined operating state. Determine the mode. Further, the ECU 2 executes various control processes such as a combustion control process and a secondary fresh air amount control process in accordance with the determined combustion mode. In the present embodiment, the ECU 2 corresponds to a control unit.

  FIG. 6 is a flowchart showing the combustion control process described above. This process is executed in synchronization with the generation of the TDC signal. In this process, first, in step 1 (illustrated as “S1”, the same applies hereinafter), it is determined whether or not the environmental condition flag F_ENV is “1”. This environmental condition flag F_ENV is set to “1” when it is determined that a temperature state suitable for combustion by self-ignition is secured in the combustion chamber 3b, and the intake air temperature TA and the engine water temperature are set. When TW is equal to or higher than each predetermined temperature, it is determined that such a temperature state is secured.

  When the determination result in step 1 is NO, assuming that the temperature state in the combustion chamber 3b suitable for self-ignition is not secured, the combustion mode is determined as the SI combustion mode, and SI combustion control is executed (step 3). Then, this process is terminated.

  On the other hand, when the determination result in step 1 is YES, it is determined whether or not the engine 3 is in an operation region (hereinafter referred to as “HCCI region”) in which CI combustion is to be performed (step 2). This determination is made according to the engine speed NE and the required torque PMCMD based on the map shown in FIG. In this map, the HCCI region is set to an operation region in which the engine speed NE is in the low to medium rotation region and the required torque PMCMD is in the low to medium load region.

  If the determination result in step 2 is NO and the engine 3 is not in the HCCI region, the combustion mode is determined to be the SI combustion mode, and after performing SI combustion control in the step 3, the present process is terminated.

  This SI combustion control is performed as follows. First, the fuel injection amount QINJ is calculated by searching a predetermined map (not shown) according to the engine speed NE and the required torque PMCMD. Next, when the engine speed NE is equal to or lower than a predetermined value and the required torque PMCMD is equal to or lower than the predetermined value, fuel of the fuel injection amount QINJ is injected from the in-cylinder fuel injection valve 10 into the cylinder C. On the other hand, at other times, a predetermined ratio (for example, 80%) of fuel is injected from the port fuel injection valve 9 to the intake port 4a with respect to the fuel injection amount QINJ, and the remaining ratio of fuel is injected from the in-cylinder fuel injection valve 10. Injection into the cylinder C. The required torque PMCMD is calculated by searching a predetermined map (not shown) according to the engine speed NE and the accelerator pedal opening AP.

  On the other hand, if the determination result in step 2 is YES and the engine 3 is in the HCCI region, the combustion mode is determined to be the CI combustion mode, and after-mentioned CI combustion control is executed (step 4), the process is terminated.

  In this combustion control process, the CI combustion flag F_STS is set to “2” in the CI combustion mode according to the determination result of the operation region of the engine 3 described above, while switching from the SI combustion mode to the CI combustion mode. Set to “1” and “0” otherwise.

  FIG. 8 shows a subroutine of this CI combustion control process. In this process, first, in step 11, the switching mechanism 85 is driven and the lever 84 and the rocker arm 64 are connected so that the exhaust valve 13 can be opened in the intake stroke.

  Next, by searching respective predetermined maps (not shown) according to the engine speed NE and the required torque PMCMD, the fuel injection amount QINJ and the fuel injection timing TINJ to be injected during the intake stroke are respectively calculated ( Steps 12 and 13), this processing is terminated. This fuel injection timing TINJ is represented by a crank angle CA. The fuel injection amount QINJ calculated as described above is a predetermined ratio from the port fuel injection valve 9 and the in-cylinder fuel injection valve 10 in the above-described SI combustion mode in the operating state other than the low rotation and low load. It is injected separately.

  FIG. 9 is a flowchart showing the secondary fresh air amount control process described above. This process is also executed in synchronization with the generation of the TDC signal. This secondary fresh air amount control process is performed for each cylinder C based on the cylinder discrimination signal. In the following, for convenience of description, the # 1 cylinder C will be described on behalf of these. In this process, first, in step 41, it is determined whether or not the above-described CI combustion flag F_STS is “2”. When the determination result is YES and the combustion mode is the CI combustion mode, the process proceeds to Step 44 described later.

  On the other hand, when the determination result of step 41 is NO, it is determined whether or not the CI combustion flag F_STS is “1” (step 42). When the determination result is NO, it is assumed that the supply of secondary fresh air is not performed, and the secondary fresh air amount GAIR2 is set to 0 (step 43), and then this process is terminated.

  On the other hand, if the decision result in the step 42 is YES and the switching from the SI combustion mode to the CI combustion mode is being performed, the process proceeds to a step 44.

  In step 44, the target temperature TEXCMD is calculated according to the engine speed NE and the required torque PMCMD. Next, it is determined whether or not the absolute value (= | TEXCMD−TEX1 |) of the difference between the target temperature TEXCMD and the # 1 exhaust temperature TEX1 is equal to or greater than a predetermined value TREF (step 45).

  As described above, once the combustion timing is deviated, an event that the temperature of the exhaust gas decreases because the combustion timing is early, and an event that the temperature of the exhaust gas increases because the combustion timing is late appear repeatedly. Therefore, when the determination result in step 45 is YES and | TEXCMD−TEX1 | ≧ TREF, it is determined that the combustion timing is greatly shifted, and the table shown in FIG. 10 is searched according to # 1 exhaust temperature TEX1. A secondary fresh air amount GAIR2 is calculated (step 46). In this table, the secondary fresh air amount GIAR2 is set to a smaller value to compensate for the higher combustion temperature in the next compression stroke as the # 1 exhaust temperature TEX1 is higher.

  Next, the secondary fresh air injection timing TAIR2 is calculated (step 47), and this process is terminated. The calculation process of the secondary fresh air injection timing TAIR2 will be described later. By opening the # 1 control valve 94a according to the secondary fresh air amount GAIR2 at the secondary fresh air injection timing TAIR2 set as described above, the secondary fresh air supplied to the # 1 exhaust port 8a is opened. The volume GAIR2 is controlled.

  For the # 2 to # 4 cylinders C, the target temperature TEXCMD calculated as described above during the CI combustion mode or switching to the CI combustion mode and # When the absolute value of the difference from 2 to # 4 exhaust temperature TEX2 to 4 is equal to or greater than a predetermined value TREF, the secondary fresh air amount GAIR2 is calculated according to # 2 to # 4 exhaust temperature TEX2 to 4.

  On the other hand, if the determination result in step 45 is NO and | TEXCMD-TEX1 | <TREF, the difference in combustion timing is relatively small, and the secondary fresh air is not supplied. The fresh air amount GAIR2 is set to the value 0, and this process is terminated.

  FIG. 11 shows a subroutine for calculating the above-described secondary fresh air injection timing TAIR2. In this process, first, in step 51, a secondary fresh air injection timing TAIR2 is calculated by searching a predetermined table (not shown) according to the secondary fresh air amount GAIR2.

  Next, it is determined whether or not the calculated secondary fresh air injection timing TAIR2 is smaller than the limit value TALMT shown in FIG. 12 (step 52). When this determination result is NO, this process is terminated as it is.

  On the other hand, if the decision result in the step 52 is YES and TAIR2 <TALMT, the limit value TALMT is set as the secondary fresh air injection timing TAIR2 (step 53), and this process is ended. As described above, the secondary fresh air injection timing TAIR2 is set between the limit value TALMT and the exhaust top dead center (a region indicated by hatching in FIG. 12), and the secondary fresh air is supplied at a timing within this region. .

  Next, the operations of the intake valve 12 and the exhaust valve 13 obtained during the CI combustion mode or during the switching to the CI combustion mode will be described together with reference to FIG. In the following description, # 1 to # 4 control valves 94a to 94d are collectively referred to as control valve 94, and # 1 to # 4 exhaust temperatures TEX1 to TEX4 are collectively referred to as exhaust temperature TEX. First, in the exhaust stroke, the exhaust valve 13 is opened by the exhaust side valve mechanism 60, and the exhaust gas is discharged to the exhaust passage 5.

  Thereafter, the control valve 94 is opened by the fresh air supply mechanism 90, whereby secondary fresh air is supplied to the exhaust passage 5, and the exhaust gas discharged into the exhaust passage 5 is cooled by the secondary fresh air. .

  In the subsequent intake stroke, the intake valve 12 is opened by the intake side valve mechanism 40 and fresh air is drawn into the cylinder C. Further, during this intake stroke, fuel is supplied from the port fuel injection valve 9 and the in-cylinder fuel injection valve 10 into the cylinder C, and this fuel is mixed with fresh air, thereby generating a mixed gas.

  Further, when the exhaust valve 13 is opened by the EGR intake mechanism 80 at the end of the intake stroke, the exhaust gas in the exhaust passage 5 is re-inhaled into the cylinder C as a high-temperature gas (hatching in the figure). The mixed gas and the high-temperature gas generated as described above are stratified into the high-temperature gas layer T1 and the mixed gas layer T2 in the cylinder C by the guide wall 3d and the protrusion 3e. The hot gas layer T1 is distributed on the exhaust valve 13 side, and the mixed gas layer T2 is distributed on the intake valve 12 side. Further, in the subsequent compression stroke, the high temperature gas layer T1 and the mixed gas layer T2 are combusted by self-ignition. This combustion is started from the higher temperature hot gas layer T1 and reaches the mixed gas layer T2.

  FIG. 15 shows an operation example obtained by the secondary fresh air control process described so far. In this example, the combustion in the SI combustion mode is performed before the timing t1, and in this state, the CI combustion flag F_STS is set to “0”. At this time, the secondary fresh air amount GAIR2 is set to a value of 0 (step 43), whereby the supply of secondary fresh air to the # 1 to # 4 exhaust ports 8a to 8d is stopped. When this state is switched to the CI combustion mode (t1), secondary fresh air is supplied to the # 1 to # 4 exhaust ports 8a to 8d (step 46). Immediately after switching to the CI combustion mode, the wall surface temperature of the cylinder C is higher due to the combustion in the SI combustion mode so far, so the secondary fresh air amount GAIR2 is set to a larger value. As a result, the exhaust gas temperature TEX and thus the high-temperature gas are cooled, and the wall surface temperature of the cylinder C is lowered accordingly. Thereby, an excessive increase in the temperature in the cylinder C (compression temperature) in the compression stroke can be avoided. Moreover, wall surface temperature falls gradually by supplying secondary fresh air.

  Thereafter, when the mode completely shifts to the CI combustion mode (t2), the CI combustion flag F_STS is switched to “2”, and the secondary fresh air control corresponding to the exhaust gas temperature TEX is continued.

  As described above, according to the present embodiment, the exhaust gas as the high-temperature gas discharged to the exhaust passage 5 is cooled by the secondary fresh air, and the secondary fresh air amount GAIR2 at this time is changed according to the exhaust temperature TEX. Since the control is performed, the temperature in the cylinder C in the subsequent compression stroke can be appropriately controlled according to the actual combustion timing, and the combustion timing in the CI combustion mode can be appropriately controlled. Thereby, the dispersion | variation in the combustion time between combustion cycles can be suppressed. Furthermore, since the secondary fresh air amount GAIR2 is controlled for each cylinder C, the variation in the combustion timing among the cylinders C can also be suppressed.

  Further, as the exhaust temperature TEX is lower and the actual combustion timing is lower, the secondary fresh air amount GAIR2 is controlled to decrease, so that the combustion temperature in the subsequent compression stroke can be increased. As a result, it is possible to avoid an event that the combustion timing is advanced or delayed, and therefore, variation in the combustion timing can be suppressed.

  Furthermore, during the switching from the SI combustion mode to the CI combustion mode, secondary fresh air is supplied and the high-temperature gas is cooled, so that an excessive increase in the temperature in the cylinder C due to self-ignition can be avoided. As a result, variation in combustion timing can be suppressed even during switching to the CI combustion mode.

  In addition, this invention can be implemented in various aspects, without being limited to the described embodiment. For example, in the embodiment, fresh air is used as the low-temperature gas, but instead, colder exhaust gas downstream from the filter 14 may be used. In the embodiment, the exhaust gas temperature TEX is used as a parameter indicating the combustion timing of fresh air and high-temperature gas in the cylinder C. Instead, another parameter indicating the combustion timing, for example, a detected cylinder is used. You may use the transition of the ionic current according to the timing which shows the peak value and maximum variation | change_quantity of the cylinder pressure calculated | required based on internal pressure, and the voltage applied to a spark plug.

  In the embodiment, the calculation of the secondary fresh air amount GAIR2 is performed only in accordance with the exhaust temperature TEX exhausted from the corresponding cylinder C. In addition to this, the secondary fresh air amount GAIR2 is exhausted from a cylinder having a 360 ° phase difference. Depending on the exhaust temperature. Specifically, when calculating the secondary fresh air amount GAIR2 sucked into the # 1 cylinder C, in addition to the # 1 exhaust temperature TEX1, according to the # 4 exhaust temperature TEX4 discharged from the # 4 cylinder C The secondary fresh air amount GAIR2 may be calculated. This is because if the temperature of the exhaust gas discharged from the # 4 cylinder C is low, the pressure of the exhaust gas also decreases, and the hot gas re-inhaled into the # 1 cylinder C decreases, so the temperature in the cylinder tends to decrease. Because. For this reason, the lower the # 4 exhaust temperature TEX4 from the # 4 cylinder C, the lower the secondary fresh air amount GIAR2 sucked into the # 1 cylinder C, thereby reducing the temperature in the # 1 cylinder C appropriately. Can be controlled.

  Furthermore, in the embodiment, the supply of secondary fresh air into the cylinder C is performed by the fresh air supply mechanism 90 via the # 1 to # 4 exhaust ports 8a to 8d. This may be performed by injecting directly into the cylinder C by the fresh air supply mechanism. Further, in the embodiment, the EGR suction mechanism 80 cannot change the valve timing of the exhaust valve 13, but may be configured to be able to change this.

  Further, the embodiment is an example in which the present invention is applied to a gasoline engine mounted on a vehicle, but the present invention is not limited to this, and may be applied to various engines such as a diesel engine other than a gasoline engine. In addition, the present invention is also applicable to engines other than those for vehicles, for example, engines for marine propulsion devices such as outboard motors having a crankshaft arranged vertically. In addition, it is possible to appropriately change the detailed configuration within the scope of the gist of the present invention.

DESCRIPTION OF SYMBOLS 1 Control apparatus 2 ECU (control means)
3 Engine 5 Exhaust passage 12 Intake valve 13 Exhaust valve 26 # 1 Exhaust temperature sensor (combustion timing parameter acquisition means)
27 # 2 exhaust temperature sensor (combustion timing parameter acquisition means)
28 # 3 exhaust temperature sensor (combustion timing parameter acquisition means)
29 # 4 exhaust temperature sensor (combustion timing parameter acquisition means)
40 Intake side valve mechanism 80 EGR intake mechanism (exhaust gas intake mechanism and low temperature gas supply mechanism)
90 New air supply mechanism (low temperature gas supply mechanism)
C cylinder TEX exhaust temperature (combustion timing parameter)
GAIR2 secondary fresh air volume

Claims (2)

As a combustion mode, it has a compression ignition combustion mode in which an air-fuel mixture is combusted by compression ignition, and a control device for an internal combustion engine that controls the internal combustion engine in the compression ignition combustion mode,
An intake side valve mechanism that draws fresh air into the cylinder by opening and closing the intake valve during the intake stroke;
An exhaust gas intake mechanism for opening and closing the exhaust valve from the end of the intake stroke to the compression stroke so that the exhaust gas discharged into the exhaust passage is again taken into the cylinder as a high-temperature gas;
A low temperature gas supply mechanism for supplying a low temperature gas having a temperature lower than that of the high temperature gas into the cylinder in order to cool the high temperature gas;
Combustion timing parameter acquisition means for acquiring a combustion timing parameter representing the combustion timing of fresh air and hot gas sucked into the cylinder;
Control means for controlling the amount of low-temperature gas supplied by the low-temperature gas supply mechanism to a decreasing side as the combustion time represented by the acquired combustion time parameter is earlier;
A control device for an internal combustion engine, comprising: The internal combustion engine further has, as the combustion mode, a spark ignition combustion mode in which an air-fuel mixture is combusted by spark ignition,
2. The internal combustion engine according to claim 1, wherein the control unit controls the amount of the low temperature gas when the combustion mode is switched from the spark ignition combustion mode to the compression ignition combustion mode. Control device.

Images