JP2014185624A - Control device of spark ignition type engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To avoid an increase in combustion noise when an operation region of performing spark ignition combustion is shifted to a region where a load is relatively higher in an operation region of performing compression ignition combustion.SOLUTION: A controller (PCM 10) interposes a first transition mode upon mode switching when an operation state of an engine body (engine 1) is shifted from an operation region of performing a spark ignition mode to a low load-side region in an operation region of performing a compression ignition mode, as well as interposes a second transition mode at the mode switching when the state is shifted from the operation region of performing the spark ignition mode to a high load-side region in the operation region of performing the compression ignition mode. The controller sets an effective compression ratio when the second transition mode is performed lower than the effective compression ratio set in a state to be shifted and lower than the effective compression ratio when the first transition mode is performed.

Description

  The technology disclosed herein relates to a control device for a spark ignition engine.

  For example, as described in Patent Document 1, a combustion mode in which an air-fuel mixture in a cylinder is ignited by compression is known as a technique for achieving both improvement in exhaust emission performance and improvement in thermal efficiency. However, the compression ignition combustion is a combustion whose pressure rises rapidly as the engine load increases, resulting in an increase in combustion noise. Therefore, as described in Patent Document 1, even in an engine that performs compression ignition combustion, spark ignition combustion is performed not by compression ignition combustion but by driving an ignition plug in an operation region on the high load side. Is common.

  In Patent Document 2, as in the engine of Patent Document 1, in an engine that performs compression ignition combustion in a low load and low rotation region, adjustment of the valve opening period of the intake and exhaust valves is performed in the region where compression ignition combustion is performed. In this way, high-temperature burned gas is retained in the cylinder, and the temperature in the cylinder is increased to promote the compression self-ignition combustion. On the other hand, in the region where the compression ignition combustion is performed, the intake valve Is described in which the burned gas in the cylinder is once blown back to the intake port side by advancing the valve opening timing, and then the burned gas is again introduced into the cylinder together with fresh air. By doing so, the temperature of the burned gas is reduced by fresh air, so that a sudden pressure increase due to compression ignition combustion is suppressed in a relatively high load high rotation region where the compression end temperature and the compression end pressure increase. The

JP 2007-154859 A JP 2009-197740 A

  By the way, since the spark ignition combustion has relatively low thermal efficiency, the combustion gas temperature becomes high. On the other hand, in compression ignition combustion, high temperature burned gas is introduced into a cylinder as described in the above-mentioned patent document in order to ensure ignitability. For this reason, in an engine that switches the combustion mode according to the engine load level, the combustion mode is changed from a high-load operation region in which spark ignition combustion is performed to a low-load operation region in which compression ignition combustion is performed. Immediately after switching from ignition combustion to compression ignition combustion, the temperature atmosphere in the cylinder is relatively high, and hot burned gas by spark ignition combustion is introduced into the cylinder, The temperature in the cylinder becomes too high. This leads to premature ignition such that the air-fuel mixture in the cylinder ignites, for example, during the compression stroke, and the pressure increase rate (dP / dθ) in the cylinder becomes steep, resulting in a large combustion noise. There is a risk of inviting.

  In particular, even in the low load side operation region where compression ignition combustion is performed, the temperature state in the cylinder can be high in the region where the load is relatively high. Therefore, even when shifting from a region where spark ignition combustion is performed to a low load side region where compression ignition combustion is performed, the region is shifted to a relatively high load region within the low load side region. In particular, a large combustion noise is likely to be caused.

  The technology disclosed herein has been made in view of such a point, and the object of the technology is from an operation region where spark ignition combustion is performed to a relatively high load region within an operation region where compression ignition combustion is performed. This is to avoid an increase in combustion noise when shifting to (1).

  In the technology disclosed herein, when switching from spark ignition combustion to compression ignition combustion, in the configuration in which the transient mode is interposed, when shifting to a relatively high load region, to a relatively low load region When the transition mode is changed between the transition mode and the transition mode to shift to a relatively high load region, the effective compression ratio is lower than the effective compression ratio set in the operation region of the transition destination. By setting the ratio, the in-cylinder temperature is lowered, so that pre-ignition of the air-fuel mixture is avoided and the generation of combustion noise is reliably avoided.

  Specifically, the technology disclosed herein relates to a spark ignition engine control device, an engine body having a cylinder, a fuel injection valve configured to inject fuel to be supplied into the cylinder, and the cylinder A spark plug configured to ignite the air-fuel mixture, and a controller configured to operate the engine body by controlling at least the fuel injection valve and the spark plug.

  When the engine body is in a predetermined operation region, the controller is in a compression ignition mode for performing compression ignition combustion in which the air-fuel mixture is combusted by self-ignition, and is in a region other than the predetermined operation region. Is a spark ignition mode in which spark ignition combustion is performed by igniting and burning the air-fuel mixture by driving the spark plug, and the controller is also in an operating region in which the operating state of the engine body performs the spark ignition mode. From the operation region in which the spark ignition mode is performed to the low-load region in the operation region in which the compression ignition mode is performed. When shifting to a region on the higher load side than the region on the low load side in the operation region to be performed, the second transient mode is interposed at the time of mode switching, The controller has an effective compression ratio at the time of execution of the second transient mode that is lower than an effective compression ratio set at the transition destination and is lower than an effective compression ratio at the time of execution of the first transient mode. Set low.

  According to this configuration, switching from the spark ignition mode to the compression ignition mode is performed when the operating state of the engine body shifts to a relatively low load side region in the operation region in which the compression ignition mode is performed, and the compression ignition mode. It is divided into a case of shifting to a relatively high load side region in the operation region where the mode is performed. Among these, when switching from the spark ignition mode according to the former transition to the compression ignition mode, the first transient mode is interposed, and when switching from the spark ignition mode according to the latter transition to the compression ignition mode. Interposes the second transient mode.

  In the second transient mode, the effective compression ratio at the time of execution is set lower than the effective compression ratio set in the transition destination. Thereby, at the time of execution of the 2nd transient mode, compression end temperature becomes low and the temperature state in a cylinder falls. In the second transient mode, compression ignition combustion may be performed. In that case, since the in-cylinder temperature is relatively low, compression ignition combustion can be stably performed without causing premature ignition. At the same time, the exhaust gas temperature to be discharged is lowered by lowering the combustion gas temperature. As a result, after performing the second transient mode, when the compression ignition mode is performed in the operation region of the transition destination, the temperature state in the cylinder is relatively low and the relatively low temperature exhaust gas is in the cylinder. Therefore, premature ignition is avoided. In this way, it is possible to avoid an increase in combustion noise when shifting from an operation region where spark ignition combustion is performed to a relatively high load region within an operation region where compression ignition combustion is performed.

  On the other hand, when shifting from the spark ignition mode operating region to the relatively low load side region in the compression ignition mode operating region, pre-ignition is relatively difficult to occur and combustion noise increase is relatively avoided. Since it is easy, it is possible to make the effective compression ratio at the time of execution of the first transient mode different from the effective compression ratio at the time of execution of the second transient mode.

  The controller may reduce the effective compression ratio by reducing intake air into the cylinder when the second transient mode is executed. In other words, in the second transient mode, the effective compression ratio may be lowered than the effective compression ratio set at the transition destination by reducing the intake air amount set at the transition destination. Adjusting the effective compression ratio by reducing the intake air makes it possible to realize the adjustment easily with good responsiveness, and is particularly effective in the transient mode at the time of mode switching.

  The controller may set the air-fuel ratio of the air-fuel mixture to lean when executing the first transient mode as compared to when executing the second transient mode.

  Unlike the second transient mode in which the in-cylinder temperature is lowered by setting the effective compression ratio low, in the first transient mode, the air-fuel ratio of the air-fuel mixture is set relatively lean. As a result, the amount of gas with respect to the amount of fuel increases, and as a result, the temperature of the exhaust gas decreases. In this way, even when shifting from the spark ignition mode to the relatively low load region in the operation region where the compression ignition mode is performed, it is avoided that the temperature state in the cylinder becomes too high immediately after switching to the compression ignition mode. The As a result, premature ignition is avoided and combustion noise is prevented from increasing. In the first transient mode, for example, the throttle valve may be fully opened to increase the amount of fresh air. By doing so, the pumping loss can be reduced in the first transient mode.

  The controller may set the EGR rate at the time of execution of the second transient mode to be lower than the EGR rate set at the transition destination.

  That is, when the second transient mode is executed, the effective compression ratio can be lowered by reducing the amount of EGR gas introduced into the cylinder, and the EGR gas having a relatively high temperature can be reduced. The temperature state inside also decreases. As a result, when shifting to a relatively high load region in the operation region where the compression ignition mode is performed, it is more reliably avoided that the in-cylinder temperature state becomes too high, thereby avoiding premature ignition. In addition, an increase in combustion noise can be avoided.

  When the controller restricts intake air in the spark ignition mode before the execution of the second transient mode, the controller may continue the intake air restriction even in the second transient mode.

  By doing so, it becomes possible to improve control responsiveness at the time of mode switching. In this case, the intake restriction that continues from the spark ignition mode to the second transient mode is released when the compression ignition mode is switched (that is, the effective compression ratio becomes relatively high).

  The controller may continue the control even in the second transient mode when the throttle ring or the intake valve is slowly closed in the spark ignition mode.

  The controller may set the fuel injection timing by the fuel injection valve after the middle of the compression stroke when the second transient mode is executed.

  As described above, when the second transient mode is executed, the in-cylinder temperature is suppressed to be low by setting the effective compression ratio low, but there is still a possibility of premature ignition. Delaying the fuel injection timing after the middle of the compression stroke is advantageous in avoiding premature ignition and, in turn, avoiding an increase in combustion noise because the reaction time of the air-fuel mixture is shortened.

  As described above, when the control device for the spark ignition engine shifts from the operation region in the spark ignition mode to the relatively high load region in the operation region in the compression ignition mode, the effective compression ratio is set at the transition destination. By setting it lower than the set effective compression ratio and lower than the effective compression ratio when shifting to the relatively low load region in the operation region of the compression ignition mode, it is particularly likely to cause an increase in combustion noise. At the time of transition, it is possible to avoid premature ignition and avoid an increase in combustion noise.

It is the schematic which shows the structure of a spark ignition type engine. It is a block diagram concerning control of a spark ignition type engine. It is sectional drawing which expands and shows a combustion chamber. It is a figure which illustrates the operating area of an engine. (A) An example of the fuel injection timing when the intake stroke injection is performed in the CI mode, an example of the heat generation rate of the CI combustion associated therewith, (b) an example of the fuel injection timing when the high pressure retarded injection is performed in the CI mode, , An example of the heat generation rate of CI combustion associated therewith, (c) an example of the fuel injection timing and ignition timing when performing high pressure retarded injection in the SI mode, and an example of the heat generation rate of SI combustion associated therewith, (d) SI 4 is an example of fuel injection timing and ignition timing when split injection of intake stroke injection and high pressure retarded injection is performed in the mode, and the heat generation rate of SI combustion associated therewith. It is a figure which compares the state of SI combustion by a high pressure retarded injection, and the state of conventional SI combustion. It is a time chart explaining the transient control at the time of switching from SI mode to relatively high load CI mode. It is a time chart explaining the transient control at the time of switching from SI mode to CI mode of relatively low load. It is a flowchart of the transient control at the time of switching from SI mode to CI mode. It is a time chart explaining the transient control different from FIG. 7 at the time of switching from the SI mode to the relatively high load CI mode. 9 is a time chart for explaining transient control different from that in FIG. 8 when switching from the SI mode to the relatively low-load CI mode.

  Hereinafter, an embodiment of a spark ignition direct injection engine will be described with reference to the drawings. The following description of preferred embodiments is exemplary. 1 and 2 show a schematic configuration of an engine (engine body) 1. The engine 1 is a spark ignition gasoline engine that is mounted on a vehicle and supplied with fuel containing at least gasoline. The engine 1 includes a cylinder block 11 provided with a plurality of cylinders 18 (only one cylinder is shown in FIG. 1, but four cylinders are provided in series, for example), and the cylinder block 11 is arranged on the cylinder block 11. The cylinder head 12 is provided, and an oil pan 13 is provided below the cylinder block 11 and stores lubricating oil. A piston 14 connected to the crankshaft 15 via a connecting rod 142 is fitted in each cylinder 18 so as to be able to reciprocate. A cavity 141 like a reentrant type in a diesel engine is formed on the top surface of the piston 14 as shown in an enlarged view in FIG. The cavity 141 is opposed to an injector 67 described later when the piston 14 is positioned near the compression top dead center. The cylinder head 12, the cylinder 18, and the piston 14 having the cavity 141 define a combustion chamber 19. The shape of the combustion chamber 19 is not limited to the illustrated shape. For example, the shape of the cavity 141, the top surface shape of the piston 14, the shape of the ceiling portion of the combustion chamber 19, and the like can be changed as appropriate.

  The engine 1 is set to a relatively high geometric compression ratio of 15 or more for the purpose of improving the theoretical thermal efficiency and stabilizing the compression ignition combustion described later. In addition, what is necessary is just to set a geometric compression ratio suitably in the range of about 15-20. As an example, the geometric compression ratio of the engine 1 is 18.

  The cylinder head 12 is provided with an intake port 16 and an exhaust port 17 for each cylinder 18. The intake port 16 and the exhaust port 17 have an intake valve 21 and an exhaust for opening and closing the opening on the combustion chamber 19 side. Each valve 22 is disposed.

  Among the valve systems that drive the intake valve 21 and the exhaust valve 22, respectively, on the exhaust side, the operation mode of the exhaust valve 22 is switched between a normal mode and a special mode, for example, a hydraulically operated variable mechanism (see FIG. 2). Hereinafter, a VVL (Variable Valve Lift) 71 is provided. Although the detailed illustration of the configuration of the VVL 71 on the exhaust side is omitted, two types of cams having different cam profiles, a first cam having one cam peak and a second cam having two cam peaks, and A lost motion mechanism is provided that selectively transmits the operating state of one of the first and second cams to the exhaust valve. When the operating state of the first cam is transmitted to the exhaust valve 22, the exhaust valve 22 operates in the normal mode in which the valve is opened only once during the exhaust stroke, whereas the operating state of the second cam is the exhaust valve. When transmitting to the engine 22, the exhaust valve 22 is operated in a special mode in which the exhaust valve 22 is opened during the exhaust stroke and is also opened during the intake stroke so that the exhaust is opened twice (FIG. 7 and the like). reference). The normal mode and special mode of the VVL 71 on the exhaust side are switched according to the operating state of the engine. Specifically, the special mode is used in the control related to the internal EGR. In the following description, operating the VVL 71 on the exhaust side in the normal mode and not opening the exhaust twice is referred to as “turning off the VVL 71”, and operating the VVL 71 on the exhaust side in the special mode to Performing the opening is sometimes referred to as “turning on the VVL 71”.

  Here, the exhaust double opening is a lift characteristic in which the exhaust valve 22 is closed in the exhaust stroke and then opened again in the intake stroke (that is, the peak of the lift curve of the exhaust valve 22 is cranked). The lift valve 22 once lifted in the exhaust stroke does not close, and the intake stroke is maintained while maintaining the predetermined opening degree. The lift characteristic (that is, the lift characteristic in which the peak of the lift curve of the exhaust valve 22 is substantially one but the base of the peak extends with the progress of the crank angle) is also included. In order to enable switching between the normal mode and the special mode, an electromagnetically driven valve system that drives the exhaust valve 22 by an electromagnetic actuator may be employed.

  Further, the execution of the internal EGR is not realized only by opening the exhaust gas twice. For example, the internal EGR control may be performed by opening the intake valve 21 twice or by opening the intake valve twice. Similarly to the exhaust double opening, the intake double opening is a lift characteristic (that is, the exhaust valve 22 of the exhaust valve 22 is reopened in the intake stroke after the intake valve 21 is substantially closed in the exhaust stroke). In addition to the lift characteristic in which the peak of the lift curve is two in line with the progress of the crank angle), the intake valve 21 once lifted in the exhaust stroke is not closed and the predetermined opening is maintained. The lift characteristic that leads to the intake stroke (that is, the lift characteristic in which the peak of the lift curve of the intake valve 21 is substantially one but the base of the peak extends in the direction opposite to the progression of the crank angle) is also included. . Further, a negative overlap period in which both the intake valve 21 and the exhaust valve 22 are closed in the exhaust stroke or the intake stroke may be provided to perform internal EGR control that causes the burned gas to remain in the cylinder 18.

  The intake side valve system is also provided with a VVL 73, similar to the exhaust side valve system provided with the VVL 71. However, unlike the VVL 71 on the exhaust side, the VVL 73 on the intake side is a cam of a large lift cam that relatively increases the lift amount of the intake valve 21 and a small lift cam that relatively decreases the lift amount of the intake valve 21. Two types of cams having different profiles, and a lost motion mechanism that selectively transmits an operating state of one of the large lift cam and the small lift cam to the intake valve 21 are configured. As shown in FIG. 7 and the like, when the operating state of the large lift cam is transmitted to the intake valve 21, the intake valve 21 opens with a relatively large lift amount, and the valve opening period also becomes longer. On the other hand, when the operating state of the small lift cam is transmitted to the intake valve 21, the intake valve 21 is opened with a relatively small lift amount, and the valve opening period is shortened. As shown in FIG. 2, a phase variable mechanism (hereinafter referred to as VVT (Variable ValveTiming)) 72 capable of changing the rotational phase of the intake camshaft with respect to the crankshaft 15 is also provided on the intake side. . The VVT 72 may employ a hydraulic, electromagnetic, or mechanical structure as appropriate, and illustration of the detailed structure is omitted. With the VVT 72 and the VVL 73, the intake valve 21 can change its valve opening timing, valve closing timing, and lift amount.

  In addition, an injector 67 that directly injects fuel into the cylinder 18 is attached to the cylinder head 12 for each cylinder 18. As shown in an enlarged view in FIG. 3, the injector 67 is disposed so that its nozzle hole faces the inside of the combustion chamber 19 from the central portion of the ceiling surface of the combustion chamber 19. The injector 67 directly injects an amount of fuel into the combustion chamber 19 at an injection timing set according to the operating state of the engine 1 and according to the operating state of the engine 1. In this example, the injector 67 is a multi-hole injector having a plurality of nozzle holes, although detailed illustration is omitted. Thereby, the injector 67 injects the fuel so that the fuel spray spreads radially from the center position of the combustion chamber 19. As indicated by the arrows in FIG. 3, the fuel spray injected radially from the central portion of the combustion chamber 19 at the timing when the piston 14 is located near the compression top dead center is a cavity formed on the top surface of the piston. It flows along the wall surface of 141. It can be paraphrased that the cavity 141 is formed so that the fuel spray injected at the timing when the piston 14 is located near the compression top dead center is contained therein. This combination of the multi-hole injector 67 and the cavity 141 is an advantageous configuration for shortening the mixture formation period and the combustion period after fuel injection. In addition, the injector 67 is not limited to a multi-hole injector, and may be an outside-opening type injector.

  A fuel tank (not shown) and the injector 67 are connected to each other by a fuel supply path. A fuel supply system 62 including a fuel pump 63 and a common rail 64 and capable of supplying fuel to the injector 67 at a relatively high fuel pressure is interposed on the fuel supply path. The fuel pump 63 pumps fuel from the fuel tank to the common rail 64, and the common rail 64 can store the pumped fuel at a relatively high fuel pressure. When the injector 67 is opened, the fuel stored in the common rail 64 is injected from the injection port of the injector 67. Here, although not shown, the fuel pump 63 is a plunger type pump and is driven by the engine 1. The fuel supply system 62 configured to include this engine-driven pump enables the fuel with a high fuel pressure of 30 MPa or more to be supplied to the injector 67. The fuel pressure may be set to about 120 MPa at the maximum. The pressure of the fuel supplied to the injector 67 is changed according to the operating state of the engine 1 as will be described later. The fuel supply system 62 is not limited to this configuration.

  As shown in FIG. 3, a spark plug 25 that ignites the air-fuel mixture in the combustion chamber 19 is attached to the cylinder head 12. In this example, the spark plug 25 is disposed through the cylinder head 12 so as to extend obliquely downward from the exhaust side of the engine 1. As shown in FIG. 3, the tip of the spark plug 25 is disposed facing the cavity 141 of the piston 14 located at the compression top dead center.

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

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

  Between the air cleaner 31 and the surge tank 33 in the intake passage 30, a water-cooled intercooler / warmer 34 that cools or heats the air and a throttle valve 36 that adjusts the amount of intake air to each cylinder 18 are arranged. It is installed. An intercooler bypass passage 35 that bypasses the intercooler / warmer 34 is also connected to the intake passage 30. The intercooler bypass passage 35 is connected to an intercooler for adjusting the flow rate of air passing through the passage 35. A bypass valve 351 is provided. Adjusting the temperature of fresh air introduced into the cylinder 18 by adjusting the ratio between the passage flow rate of the intercooler bypass passage 35 and the passage flow rate of the intercooler / warmer 34 through the opening degree adjustment of the intercooler bypass valve 351. Is possible.

  The upstream portion of the exhaust passage 40 is constituted by an exhaust manifold having an independent passage branched for each cylinder 18 and connected to the outer end of the exhaust port 17 and a collecting portion where the independent passages gather. A direct catalyst 41 and an underfoot catalyst 42 are connected downstream of the exhaust manifold in the exhaust passage 40 as exhaust purification devices for purifying harmful components in the exhaust gas. Each of the direct catalyst 41 and the underfoot catalyst 42 includes a cylindrical case and, for example, a three-way catalyst disposed in a flow path in the case.

  A portion between the surge tank 33 and the throttle valve 36 in the intake passage 30 and a portion upstream of the direct catalyst 41 in the exhaust passage 40 are used for returning a part of the exhaust gas to the intake passage 30. They are connected via a passage 50. The EGR passage 50 includes a main passage 51 in which an EGR cooler 52 for cooling the exhaust gas with engine coolant is disposed, and an EGR cooler bypass passage 53 for bypassing the EGR cooler 52. ing. The main passage 51 is provided with an EGR valve 511 for adjusting the amount of exhaust gas recirculated to the intake passage 30, and the EGR cooler bypass passage 53 has a flow rate of exhaust gas flowing through the EGR cooler bypass passage 53. An EGR cooler bypass valve 531 for adjustment is provided.

  The engine 1 configured as described above is controlled by a powertrain control module (hereinafter referred to as PCM) 10. The PCM 10 includes a microprocessor having a CPU, a memory, a counter timer group, an interface, and a path connecting these units. This PCM 10 constitutes a controller.

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

  The PCM 10 determines the state of the engine 1 and the vehicle by performing various calculations based on these detection signals, and accordingly, the injector 67, the spark plug 25, the VVT 72, the intake side VVL 73, the exhaust side VVL 71, Control signals are output to the fuel supply system 62 and actuators of various valves (throttle valve 36, intercooler bypass valve 351, EGR valve 511, and EGR cooler bypass valve 531). Thus, the PCM 10 operates the engine 1.

  FIG. 4 shows an example of the operation region when the engine 1 is warm. This engine 1 is a compression ignition combustion in which combustion is performed by compression self-ignition without ignition by the spark plug 25 in a low load region where the engine load is relatively low for the purpose of improving fuel consumption and exhaust emission performance. I do. However, as the load on the engine 1 increases, in the compression ignition combustion, the combustion becomes too steep and causes problems such as combustion noise. Therefore, in the engine 1, in a high load region where the engine load is relatively high, the compression ignition combustion is stopped, and the engine 1 is switched to the spark ignition combustion using the spark plug 25. As described above, the engine 1 has a CI (Compression Ignition) mode in which compression ignition combustion is performed and an SI (Spark Ignition) mode in which spark ignition combustion is performed in accordance with the operation state of the engine 1, in particular, the load of the engine 1. It is configured to switch. The mode switching boundary line is not limited to the illustrated example. In addition to the load level, the engine 1 is configured to switch modes according to the operating state under various conditions as described later.

  The CI mode is further divided into two areas depending on the engine load. Specifically, in the low to medium load region (1) in the CI mode, hot EGR gas having a relatively high temperature is introduced into the cylinder 18 in order to improve the ignitability and stability of the compression ignition combustion. This is because the VVL 71 on the exhaust side is turned on and the exhaust valve 22 is opened twice during the intake stroke. The introduction of hot EGR gas is advantageous in increasing the compression end temperature in the cylinder 18 and improving the ignitability and stability of the compression ignition combustion in the region (1). In the region (1), as shown in FIG. 5A, the injector 67 injects fuel into the cylinder 18 at least during the period from the intake stroke to the middle of the compression stroke, thereby forming a homogeneous mixture. To do. In the region (1), the air-fuel ratio of the air-fuel mixture is basically set to the stoichiometric air-fuel ratio (A / F = 14.7 ± 0.5, excess air ratio λ≈1). However, as indicated by the one-dot chain line in FIG. 4, in the region (1), the air-fuel ratio of the air-fuel mixture is set to be leaner than the stoichiometric air-fuel ratio in a relatively low load and low speed partial region.

  Thus, in this region (1), the air-fuel mixture in the combustion chamber 19 undergoes compression self-ignition near the compression top dead center as shown in FIG.

  In the CI mode, in the region (2) where the load is higher than the region (1), the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1). By setting the stoichiometric air-fuel ratio, the three-way catalyst can be used and, as will be described later, the air-fuel ratio of the air-fuel mixture is also set to the stoichiometric air-fuel ratio in the SI mode, so that switching between the SI mode and the CI mode is possible. The control at the time is simplified, and further, the CI mode can be expanded to the high load side.

  In the region (2), the temperature in the cylinder 18 naturally increases as the engine load increases, so the hot EGR gas amount is reduced to avoid pre-ignition. This is due to the adjustment of the amount of internal EGR gas introduced into the cylinder 18. Further, the hot EGR gas amount may be adjusted by adjusting the external EGR gas amount bypassing the EGR cooler 52.

  In the region (2), a cooled EGR gas having a relatively low temperature is further introduced into the cylinder 18. Thus, by introducing the hot hot EGR gas and the cold cooled EGR gas into the cylinder 18 at an appropriate ratio, the compression end temperature in the cylinder 18 is made appropriate, and the rapid ignition while ensuring the ignitability of the compression ignition. Avoid combustion and stabilize compression ignition combustion.

  In this way, in the region (2) including the switching boundary between the CI mode and the SI mode, the temperature in the cylinder 18 is lowered, but the compression end temperature in the cylinder 18 is still higher. obtain. If the fuel is injected into the cylinder 18 within the period from the intake stroke to the middle of the compression stroke as in the region (1), abnormal combustion such as pre-ignition may occur. On the other hand, if a large amount of cooled EGR gas having a low temperature is introduced to lower the compression end temperature in the cylinder, the ignitability of the compression ignition is deteriorated. That is, since compression ignition combustion cannot be stably performed only by temperature control in the cylinder 18, in this region (2), in addition to temperature control in the cylinder 18, it is prematurely devised by devising a fuel injection mode. Stabilize compression ignition combustion while avoiding abnormal combustion such as ignition. Specifically, this fuel injection mode has a fuel pressure significantly higher than that in the conventional case, and as shown in FIG. 5 (b), at least a period from the latter stage of the compression stroke to the initial stage of the expansion stroke (hereinafter referred to as this The fuel is injected into the cylinder 18 within a period (referred to as a retard period). This characteristic fuel injection mode is hereinafter referred to as “high pressure retarded injection” or simply “retarded injection”. By such high-pressure retarded injection, compression ignition combustion is stabilized while avoiding abnormal combustion in the region (2). Details of the high-pressure retarded injection will be described later.

  In contrast to the CI mode divided into two areas according to the engine load level, the SI mode is roughly divided into two areas (3) and (4) according to the engine speed. It has been. The region (3) corresponds to a low speed region when the operation region of the engine 1 is divided into two, a low speed and a high speed, and a part on the low load side of the high speed region in the illustrated example. Corresponds to part of the high load side of the high-speed range The boundary between the region (3) and the region (4) is not limited to the illustrated example.

  In each of the region (3) and the region (4), the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1) as in the region (2). Therefore, the air-fuel ratio of the air-fuel mixture is made constant at the theoretical air-fuel ratio (λ≈1) across the boundary between the CI mode and the SI mode. Further, in the SI mode (that is, in the region (3) and the region (4)), basically, the throttle valve 36 is fully opened while the opening of the EGR valve 511 is adjusted to be introduced into the cylinder 18. Adjust fresh air volume and external EGR gas volume. Even in the SI mode region, the throttle valve 36 may be throttled in a relatively low load region. Adjusting the ratio of gas introduced into the cylinder 18 reduces pump loss and introduces a large amount of EGR gas into the cylinder 18 to reduce the combustion temperature of spark ignition combustion and reduce cooling loss. It is done. In the SI mode region, external EGR gas cooled mainly through the EGR cooler 52 is introduced into the cylinder 18. This is advantageous for avoiding abnormal combustion and also has an advantage of suppressing generation of Raw NOx. In the fully open load range, the external EGR is set to zero by closing the EGR valve 511.

  In the SI mode region, instead of introducing EGR gas, the introduction is stopped, while the opening degree of the throttle valve 36 is controlled in accordance with the amount of fuel to be injected. The amount of fresh air introduced into the cylinder 18 may be adjusted so that ≈1).

  As described above, the geometric compression ratio of the engine 1 is set to 15 or more (for example, 18). Since the high compression ratio increases the compression end temperature and the compression end pressure, it is advantageous for stabilizing the compression ignition combustion in the CI mode, particularly in a low load region (for example, the region (1)). On the other hand, the high compression ratio engine 1 has a problem that abnormal combustion such as pre-ignition and knocking is likely to occur in the SI mode which is a high load region.

  Therefore, in the engine 1, in the SI mode region (3) and region (4), abnormal combustion is avoided by performing the above-described high-pressure retarded injection. More specifically, in the region (3), as shown in FIG. 5 (c), fuel is injected into the cylinder 18 in the retard period from the late stage of the compression stroke to the early stage of the expansion stroke with a high fuel pressure of 30 MPa or more. Only high-pressure retarded injection is performed. On the other hand, in the region (3), as shown in FIG. 5D, a part of the fuel to be injected is injected into the cylinder 18 within the intake stroke period in which the intake valve 21 is open. The remaining fuel is injected into the cylinder 18 within the retard period. That is, in region (4), fuel split injection is performed. Here, the intake stroke period during which the intake valve 21 is open is not a period defined based on the piston position, but a period defined based on the opening / closing of the intake valve, and the intake stroke referred to here is VVL73. Depending on the closing timing of the intake valve 21 changed by the VVT 72, the piston may deviate from the time when the piston reaches the bottom dead center of the intake.

  Next, the high pressure retarded injection in the SI mode will be described with reference to FIG. FIG. 6 shows the heat generation rate (upper diagram) and the progress of the unburned mixture reaction in the SI combustion (solid line) by the high pressure retarded injection described above and the conventional SI combustion (broken line) in which fuel injection is performed during the intake stroke. It is a figure which compares the difference in a degree (lower figure). The horizontal axis in FIG. 6 is the crank angle. As a premise of this comparison, the operating state of the engine 1 is both a high load low speed region (that is, region (3)), and the amount of fuel to be injected is the case of SI combustion by high pressure retarded injection and conventional SI combustion. They are the same as each other.

  First, in the conventional SI combustion, a predetermined amount of fuel is injected into the cylinder 18 during the intake stroke (broken line in the upper diagram). In the cylinder 18, a relatively homogeneous air-fuel mixture is formed after the fuel injection until the piston 14 reaches the compression top dead center. In this example, ignition is executed at a predetermined timing indicated by a white circle after the compression top dead center, thereby starting combustion. After the start of combustion, as shown by the broken line in the upper diagram of FIG. 6, the combustion ends through a peak of the heat generation rate. The time from the start of fuel injection to the end of combustion corresponds to the reaction possible time of the unburned mixture (hereinafter sometimes simply referred to as reaction possible time). As shown by the broken line in the lower diagram of FIG. The reaction of the fuel mixture gradually proceeds. The dotted line in the figure shows the ignition threshold, which is the reactivity with which the unburned mixture reaches ignition, and the conventional SI combustion has a very low reaction time in combination with the low speed range. In the meantime, the reaction of the unburned mixture continues to progress during that time, so the reactivity of the unburned mixture exceeds the ignition threshold before and after ignition, and abnormal combustion such as premature ignition or knocking occurs. cause.

  On the other hand, the high pressure retarded injection aims to shorten the reaction possible time, thereby avoiding abnormal combustion. That is, as shown in FIG. 6, the possible reaction time is a period during which the injector 67 injects fuel ((1) injection period), and after the injection is completed, a combustible mixture is formed around the spark plug 25. (2) The mixture formation period) and the period until the combustion started by ignition is completed ((3) combustion period), that is, (1) + (2 ) + (3). The high-pressure retarded injection shortens the injection period, the mixture formation period, and the combustion period, thereby shortening the reaction time. This will be described in order.

  First, the high fuel pressure relatively increases the amount of fuel injected from the injector 67 per unit time. For this reason, when the fuel injection amount is constant, the relationship between the fuel pressure and the fuel injection period is generally such that the lower the fuel pressure, the longer the injection period, and the higher the fuel pressure, the shorter the injection period. Therefore, the high pressure retarded injection in which the fuel pressure is set to be significantly higher than the conventional one shortens the injection period.

  Further, the high fuel pressure is advantageous for atomization of the fuel spray injected into the cylinder 18 and makes the flight distance of the fuel spray longer. For this reason, the relationship between the fuel pressure and the fuel evaporation time is generally longer as the fuel pressure is lower, and the fuel evaporation time is longer as the fuel pressure is higher. Further, the time until the fuel spray reaches the fuel pressure and the spark plug 25 is generally longer as the fuel pressure is lower, and the time until the fuel spray is higher as the fuel pressure is higher. The air-fuel mixture formation period is a time obtained by adding the fuel evaporation time and the fuel spray arrival time around the spark plug 25. Therefore, the higher the fuel pressure, the shorter the air-fuel mixture formation period. Therefore, the high pressure retarded injection in which the fuel pressure is set to be significantly higher than the conventional case shortens the mixture formation period as a result of the fuel evaporation time and the fuel spray arrival time around the spark plug 25 being reduced. On the other hand, as shown by white circles in the figure, the conventional intake stroke injection at a low fuel pressure significantly increases the mixture formation period. In the SI mode, the combination of the multi-injector type injector 67 and the cavity 141 shortens the time until fuel spray reaches around the spark plug 25 after fuel injection. Effective for shortening.

  Thus, shortening the injection period and the mixture formation period makes it possible to set the fuel injection timing, more precisely, the injection start timing to a relatively late timing. Therefore, in the high pressure retarded injection, as shown in the upper diagram of FIG. 6, fuel injection is performed within the retard period from the latter stage of the compression stroke to the early stage of the expansion stroke. As the fuel is injected into the cylinder 18 at a high fuel pressure, the turbulence in the cylinder becomes stronger and the turbulence energy in the cylinder 18 increases. This high turbulence energy is a timing at which the fuel injection timing is relatively late. Therefore, it is advantageous for shortening the combustion period.

  That is, when the fuel injection is performed within the retard period, the relationship between the fuel pressure and the turbulent energy in the combustion period is generally lower as the fuel pressure is lower and the turbulent energy is lower as the fuel pressure is higher. Becomes higher. Here, even if fuel is injected into the cylinder 18 at a high fuel pressure, if the injection timing is in the intake stroke, the time until the ignition timing is long, or the cylinder 18 is in the compression stroke after the intake stroke. Due to the compression of the inside, the disturbance in the cylinder 18 is attenuated. As a result, when fuel is injected during the intake stroke, the turbulent energy during the combustion period becomes relatively low regardless of the fuel pressure level.

  In general, the relationship between the turbulent energy and the combustion period in the combustion period is such that the lower the turbulent energy, the longer the combustion period, and the higher the turbulent energy, the shorter the combustion period. Therefore, the relationship between the fuel pressure and the combustion period is such that the lower the fuel pressure, the longer the combustion period, and the higher the fuel pressure, the shorter the combustion period. That is, the high pressure retarded injection shortens the combustion period. In contrast, the conventional intake stroke injection at a low fuel pressure has a longer combustion period. The multi-injector type injector 67 is advantageous for improving the turbulent energy in the cylinder 18 and is effective for shortening the combustion period. In addition, the combination of the multi-injector type injector 67 and the cavity 141 provides fuel. Putting the spray in the cavity 141 is also effective for shortening the combustion period.

  As described above, the high pressure retarded injection shortens the injection period, the mixture formation period, and the combustion period, and as a result, as shown in FIG. 6, the fuel injection start timing SOI to the combustion end timing θend are not changed. It is possible to significantly shorten the reaction time of the fuel mixture as compared with the case of fuel injection during the conventional intake stroke. As a result of shortening this reaction possible time, as shown in the upper diagram of FIG. 6, in the conventional intake stroke injection at a low fuel pressure, as shown by a white circle, the reaction progress of the unburned mixture at the end of combustion is shown. However, when the ignition threshold is exceeded and abnormal combustion occurs, the high-pressure retarded injection suppresses the progress of the reaction of the unburned mixture at the end of combustion, as shown by the black circle, to prevent abnormal combustion. It can be avoided. It should be noted that the ignition timing is set to the same timing in the white circle and the black circle in the upper diagram of FIG.

  By setting the fuel pressure to, for example, 30 MPa or more, the combustion period can be effectively shortened. Moreover, the fuel pressure of 30 MPa or more can effectively shorten the injection period and the mixture formation period, respectively. The fuel pressure is preferably set as appropriate according to the properties of the fuel used, which contains at least gasoline. The upper limit may be 120 MPa as an example.

  The high pressure retarded injection avoids the occurrence of abnormal combustion in the SI mode by devising the form of fuel injection into the cylinder 18. Unlike this, it is conventionally known that the ignition timing is retarded for the purpose of avoiding abnormal combustion. The retarding of the ignition timing suppresses the progress of the reaction by suppressing the increase in the temperature and pressure of the unburned mixture. However, retarding the ignition timing leads to a decrease in thermal efficiency and torque, whereas when performing high-pressure retarded injection, the ignition timing can be advanced by an amount that avoids abnormal combustion by devising the form of fuel injection. Since it is possible, thermal efficiency and torque are improved. That is, the high pressure retarded injection not only avoids abnormal combustion, but also makes it possible to advance the ignition timing by the amount that can be avoided, which is advantageous in improving fuel consumption.

  As described above, the high pressure retarded injection in the SI mode can shorten the injection period, the mixture formation period, and the combustion period, but the high pressure retarded injection performed in the CI mode region (2) It is possible to shorten the injection period and the mixture formation period. In other words, the turbulence in the cylinder 18 is increased by injecting the fuel into the cylinder 18 at a high fuel pressure, so that the mixing performance of the atomized fuel is increased and the fuel is injected at a late timing near the compression top dead center. However, a relatively homogeneous air-fuel mixture can be quickly formed.

  In the high pressure retarded injection in the CI mode, fuel is injected at a late timing near the compression top dead center in a relatively high load region, for example, while preventing premature ignition during the compression stroke period, as described above. Since a substantially homogeneous air-fuel mixture is quickly formed, it is possible to reliably perform compression ignition after the compression top dead center. Thus, in the expansion stroke period in which the pressure in the cylinder 18 gradually decreases due to motoring, the compression ignition combustion is performed, so that the combustion becomes slow, and the pressure increase in the cylinder 18 due to the compression ignition combustion (dP / It is avoided that dθ) becomes steep. Thus, as a result of eliminating the NVH restriction, the CI mode region is expanded to the high load side.

  Returning to the description of the SI mode, as described above, the high pressure retarded injection in the SI mode shortens the reaction time of the unburned mixture by performing the fuel injection within the retard period. In the low speed range where the engine 1 has a relatively low rotational speed, the actual time for the crank angle change is long, which is effective. On the other hand, in the high speed range where the engine 1 has a relatively high rotational speed, the actual Not very effective due to short time. On the contrary, in the retard injection, since the fuel injection timing is set near the compression top dead center, in-cylinder gas not containing fuel, in other words, air having a high specific heat ratio is compressed in the compression stroke. As a result, in the high speed region, the compression end temperature in the cylinder 18 becomes high, and this high compression end temperature may cause knocking. Therefore, when only retarded injection is performed in the high load and high speed region (4) where the amount of fuel to be injected increases, the ignition timing may be retarded to avoid knocking. .

  Therefore, as shown in FIG. 4, in the region (4) where the rotational speed is relatively high and the load is high in the SI mode, as shown in FIG. The fuel is injected into the cylinder 18 within the period, and the remaining fuel is injected into the cylinder 18 within the retard period. In the intake stroke injection, it is possible to lower the specific heat ratio of the in-cylinder gas (that is, the air-fuel mixture containing fuel) during the compression stroke, thereby keeping the compression end temperature low. Thus, since the compression end temperature is lowered, knocking can be suppressed, so that the ignition timing can be advanced.

  Further, by performing the high pressure retarded injection, as described above, the turbulence becomes strong in the cylinder 18 (combustion chamber 19) near the compression top dead center, and the combustion period is shortened. This is also advantageous in suppressing knocking, and the ignition timing can be further advanced. Thus, in the region (4), by performing split injection of intake stroke injection and high pressure retarded injection, it is possible to improve thermal efficiency while avoiding abnormal combustion.

  In order to shorten the combustion period in the region (4), a multi-point ignition configuration may be adopted instead of performing the high pressure retarded injection. That is, a plurality of ignition plugs are arranged facing the combustion chamber, and in the region (4), the intake stroke injection is executed and each of the plurality of ignition plugs is driven to perform multipoint ignition. By doing so, since the flame spreads from each of the plurality of fire types in the combustion chamber 19, the flame spreads quickly and the combustion period is shortened. As a result, the combustion period is shortened similarly to the case where high pressure retarded injection is employed, which is advantageous for improving the thermal efficiency.

(Control when switching from SI mode to CI mode)
Since spark ignition combustion has lower thermal efficiency than compression ignition combustion, the combustion gas temperature becomes relatively high. On the other hand, in the CI mode in which compression ignition combustion is performed, as described above, in order to ensure the ignitability of compression ignition, at least the internal EGR gas is introduced into the cylinder 18 to increase the temperature state in the cylinder 18. doing.

  Immediately after switching from the SI mode where the combustion gas temperature is relatively high to the CI mode, the inside of the cylinder 18 is in a high temperature atmosphere, and exhaust gas having a high temperature generated by the spark ignition combustion is introduced into the cylinder 18. Thus, compression ignition combustion is performed in a state where the temperature state in the cylinder 18 is high. In this case, for example, if fuel is injected into the cylinder 18 at a relatively early time such as during the intake stroke, pre-ignition occurs during the compression stroke period, and the pressure increase rate (dP / dθ in the cylinder 18). ) May become steep and large combustion noise may occur.

  Therefore, in this engine 1, transient control is performed to avoid premature ignition when switching from the SI mode to the CI mode and to avoid an increase in combustion noise.

  Switching from the SI mode to the CI mode is, for example, when the load of the engine 1 shifts from the high load region in the SI mode to the low load region in the CI mode in the warm operation map shown in FIG. Can correspond to That is, the SI mode is switched to the CI mode as the load on the engine 1 is reduced.

  Here, as described above, even in the low load region in which the CI mode is set, as described above, in the relatively high load region (2), the temperature state in the cylinder is relatively high, and during the steady operation. However, it is easy to cause premature ignition. On the other hand, in the relatively low load region (1), pre-ignition is unlikely to occur compared to the region (2). Therefore, even when the mode is switched from the SI mode to the CI mode, the transition destination of the operating state of the engine 1 is in the relatively high load region (2) (see the arrow (a) in FIG. 4). As compared with the case of shifting to the relatively low load region (1) (see the arrow in FIG. 4B), premature ignition is likely to occur, and even more countermeasures are required for transient control.

  Therefore, in the engine 1, when switching from the SI mode to the CI mode, the transient control (that is, the first control) when the operation state transition destination of the engine 1 is the relatively low load region (1). (Transient mode) and transient control (that is, second transient mode) in the case of a relatively high load region (2) are different from each other.

  FIG. 7 shows a time chart of the transient control at the time of switching from the SI mode to the region (2) in the CI mode. Specifically, FIG. 7 shows the fuel injection at the time of switching from the SI mode to the CI mode. An example of a change in timing and spark ignition timing, a change in in-cylinder pressure, a change in opening state of an intake / exhaust valve, a change in opening of a throttle valve, and a change in gas state in a cylinder are shown. In FIG. 7, the crank angle (that is, time) advances from the left to the right of the page. Note that the changes in the fuel injection timing, spark ignition timing, and in-cylinder pressure shown in FIG. 7 are examples in the description of the invention, and are not limited to the illustrated timing (the same applies to FIG. 8 and the like). is there).

  First, in the first cycle on the leftmost side in FIG. 7, the operation is performed in the high-load SI mode. Here, the fuel injection is executed within the period from the latter half of the compression stroke to the early stage of the expansion stroke (that is, the high pressure retarded injection). ) And spark ignition is performed near the compression top dead center. The air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1), and the intake-side VVL 73 drives the intake valve 21 with a large lift cam so that the amount of fresh air is commensurate with the fuel injection amount. In addition, the VVT 72 sets the valve closing timing to a later timing after the intake bottom dead center. In this way, the intake valve 21 is closed late to limit the amount of fresh air (see also the gas state in the cylinder shown in the lowermost stage of FIG. 7). In the illustrated example, in the first cycle, the amount of fresh air that is not limited only by the control of the intake valve 21 is compensated by the throttle valve 36 being throttled. Further, the external EGR gas is introduced into the cylinder 18 by opening the EGR valve 511 and / or the EGR cooler bypass valve 531. The VVL 71 on the exhaust side is off and the internal EGR gas is not introduced. Thus, in the first cycle in which spark ignition combustion is executed, the exhaust gas temperature becomes high (high-temperature burned gas).

  The second cycle corresponds to a cycle at the time of switching from the SI mode to the CI mode, and corresponds to the second transient mode. In this second transient mode, the effective compression ratio is lowered, thereby lowering the temperature state in the cylinder. That is, in the second cycle, the throttle opening is reduced, and the VVL 73 on the intake side switches from the large lift cam to the small lift cam and sets the valve closing timing after the intake bottom dead center. As a result, the amount of fresh air introduced into the cylinder 18 is reduced. Further, the EGR valve 511 and / or the EGR cooler bypass valve 531 remain open, and external EGR gas is introduced into the cylinder 18. Further, the exhaust-side VVL 71 also remains off, and the internal EGR gas is not introduced into the cylinder 18. As a result, the effective compression ratio is lower than in the first cycle, and lower than in the third cycle described later.

  Although the amount of fuel injected by the injector 67 is set to be substantially the same as in the first cycle, the injection timing is set after the middle stage of the compression stroke.

  Thus, in the second cycle, the spark plug 25 is not operated, and the air-fuel mixture is compressed and ignited. As described above, by reducing the effective compression ratio by limiting the amount of fresh air, the compression end temperature and the compression end pressure decrease, but the temperature in the cylinder is relatively high due to spark ignition combustion up to one cycle. At the same time, the high-temperature burned gas is introduced into the cylinder 18 as the external EGR gas, and the retarded injection with the fuel injection timing delayed is used so that the air-fuel mixture is surely self-ignited near the compression top dead center. And it comes to burn stably.

  Thus, in the second cycle, by performing compression ignition combustion with the effective compression ratio lowered, the temperature state in the cylinder is lowered and the exhaust gas temperature discharged after the second cycle combustion is also lowered. To come. In the gas state shown at the bottom of FIG. 7, the temperature level of the “burned gas” is shown by the size of the hatched pitch width. The narrow pitch width means that the burned gas temperature is high. Correspondingly, a wide pitch width corresponds to a low temperature of burned gas.

  The subsequent third cycle corresponds to a cycle immediately after switching from the SI mode to the CI mode (that is, after transition to the region (2)). In the third cycle, the VVT 72 is operated to advance the closing timing of the intake valve 21 more than in the second cycle, and the exhaust-side VVL 71 is turned on to open the exhaust twice. Thereby, although a part of the burned gas generated by the compression ignition combustion in the second cycle is introduced into the cylinder 18, as described above, the temperature of the burned gas is suppressed to be low and 2 In the cycle, the temperature state in the cylinder is also kept low, so the temperature state in the cylinder 18 in the third cycle is not so high. Further, in the third cycle, while the throttle valve 36 that has been throttled in the second cycle is fully opened, the EGR valve 511 and / or the EGR cooler bypass valve 531 are continuously opened, so that the external EGR gas is released. It is introduced into the cylinder 18. In this way, the intake air (gas amount) introduced into the cylinder 18 in the third cycle increases compared to the second cycle, so the effective compression ratio in the third cycle is higher than the effective compression ratio in the second cycle. Become. Further, the EGR rate in the third cycle is higher than the EGR rate in the second cycle by the amount of internal EGR gas introduced. In other words, in the second cycle, the EGR rate is lower than that in the third cycle, and the amount of burned gas in the cylinder 18 is reduced, which is advantageous for lowering the in-cylinder temperature. Is possible.

  The amount of fuel injected by the injector 67 is about the same as in the second cycle, and the A / F of the air-fuel mixture is set to the stoichiometric air-fuel ratio. Further, as set in the region (2), the fuel injection timing is within the period from the latter half of the compression stroke to the initial stage of the expansion stroke (that is, retard injection). The A / F of the air-fuel mixture in the third cycle may be set leaner than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio depending on the operating state of the engine 1.

  In this third cycle, the spark plug 25 is deactivated as in the second cycle. As described above, in the third cycle, the temperature state in the cylinder 18 does not become excessively high, and by performing the retard injection, the compression ignition is reliably performed in the vicinity of the compression top dead center without premature ignition. Burns stably. In this way, it is avoided that combustion noise increases immediately after switching from the high load SI mode to the relatively high load region (2) in the CI mode. From the third cycle onward, when the switching from the SI mode to the CI mode is completed, combustion control according to the operating state of the engine 1 is executed.

  FIG. 8 shows a time chart of the transient control at the time of switching from the SI mode to the region (1) in the CI mode. That is, in the first cycle on the leftmost side in FIG. 8, the operation is performed in the high load SI mode. Here, fuel injection is performed during the compression stroke period (that is, high pressure retarded injection) and compression top dead. Spark ignition is performed near the point. The air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1), and the intake-side VVL 73 drives the intake valve 21 with a large lift cam so that the amount of fresh air is commensurate with the fuel injection amount. In addition, the VVT 72 sets the valve closing timing to a later timing after the intake bottom dead center. Thus, the fresh air quantity is limited by closing the intake valve 21 late. In the first cycle, the throttle valve 36 is throttled and the throttle valve 36 is gradually opened in the fully open direction in preparation for the transition to the region (1) where the throttle valve 36 is fully opened. Further, the point that the external EGR gas is introduced into the cylinder 18 by opening the EGR valve 511 and / or the EGR cooler bypass valve 531 in the first cycle is the same as the time chart of FIG. The VVL 71 on the exhaust side is off and the internal EGR gas is not introduced. In the first cycle in which spark ignition combustion is performed, the exhaust gas temperature can be high.

  In the subsequent second cycle (that is, the first transient mode corresponding to the cycle at the time of switching from the SI mode to the CI mode), unlike FIG. 7, the throttle opening is set to fully open, while the VVL 73 on the intake side is set. Switches from a large lift cam to a small lift cam. At this time, the phase of the intake valve 21 does not change, whereby the closing timing of the intake valve 21 is instantaneously switched to the vicinity of the intake bottom dead center. As a result, the amount of fresh air introduced into the cylinder 18 increases. Note that the opening / closing timing of the intake valve 21 corresponds to an exhaust double opening described later. Further, since the external EGR gas is not introduced after the transition to the region (1), the EGR valve 511 and the EGR cooler bypass valve 531 are fully closed, and the introduction of the external EGR gas into the cylinder 18 is stopped. However, the external EGR has low control responsiveness, and even after the EGR valve 511 and the EGR cooler bypass valve 531 are fully closed, the relatively high-temperature exhaust gas remaining in the EGR passage 50 remains in the cylinder 18 in the second cycle. (See the gas state at the bottom of FIG. 8). Thus, in the first transient mode when shifting to the region (1), compared to the second transient mode shifting to the region (2), the intake amount (gas amount) into the cylinder 18 in the transient mode. And the effective compression ratio becomes relatively high. Further, the pumping loss can be reduced in the first transient mode.

  The amount of fuel injected by the injector 67 is set to be almost the same as that in the first cycle. As a result, in the second cycle, the amount of new air increases, and the A / F of the mixture is mixed in the first cycle. It is set to be lean as compared with the A / F (that is, the stoichiometric air-fuel ratio). Further, the A / F becomes lean even when compared with the second transient mode when shifting to the region (2). The fuel injection timing is set during the intake stroke as illustrated in FIG. 8, thereby forming a relatively homogeneous lean air-fuel mixture.

  Thus, in the second cycle, the spark plug 25 is not operated, and the air-fuel mixture leaner than the stoichiometric air-fuel ratio is compressed and ignited in the vicinity of the compression top dead center. By making the air-fuel mixture lean, the amount of gas with respect to the amount of fuel increases, which is advantageous for lowering the combustion gas temperature, and by performing compression ignition combustion, the combustion gas temperature also decreases compared to spark ignition combustion Therefore, the exhaust gas temperature exhausted after the second cycle combustion is significantly lowered.

  In the subsequent third cycle, the exhaust VVL 71 is turned on to open the exhaust twice. Thereby, although a part of the burned gas generated by the compression ignition combustion in the second cycle is introduced into the cylinder 18, as described above, the temperature of the burned gas is suppressed to be low and 2 In the cycle, the temperature state in the cylinder is also kept low, so the temperature state in the cylinder 18 in the third cycle is not so high. Further, in the third cycle, as in the second cycle, the intake valve 21 remains in a small lift and the throttle valve 36 is set to fully open, so that the internal EGR gas flows into the cylinder 18 as shown in FIG. The amount of fresh air is reduced by the amount introduced into the air. As a result, the A / F of the air-fuel mixture in the third cycle becomes richer than the A / F of the air-fuel mixture in the second cycle. The A / F of the air-fuel mixture in the third cycle may be set leaner than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio depending on the operating state of the engine 1. In the third cycle, fuel injection is set during the intake stroke.

  In this third cycle, the spark plug 25 is deactivated as in the second cycle. As described above, in the third cycle, the temperature state in the cylinder 18 does not become excessively high, so that the relatively homogeneous mixture formed in the cylinder 18 is excessively injected by injecting fuel during the intake stroke. Without igniting prematurely, it is reliably ignited near the compression top dead center and burns stably. Thus, an increase in combustion noise immediately after switching from the SI mode to the relatively low load region (1) in the CI mode is avoided.

  Thus, when shifting from the SI mode to the relatively high load area (2) in the CI mode, the effective compression ratio is set lower than the effective compression ratio set in the operation area of the transfer destination. By performing compression ignition combustion, the temperature state in the cylinder is prevented from becoming high. On the other hand, when shifting from the SI mode to the relatively low load region (1) in the CI mode, By performing compression ignition combustion with the A / F being leaner than the stoichiometric air-fuel ratio, the exhaust gas temperature is lowered, and this increases combustion noise when switching from the SI mode to the CI mode. Is effectively avoided.

  Here, in the relatively light load region in the SI mode, the introduction of intake air (fresh air) may be limited by throttling or slow closing control of the intake valve. When shifting from such a state to the region (2) in the CI mode, the throttling in the SI mode and the slow closing control of the intake valve may be continued as they are in the second transient mode. By doing this, in the second transient mode, the effective compression ratio can be set lower than the effective compression ratio set in the transition destination, and there is an advantage that controllability is improved.

  Next, a control flow executed by the PCM 10 for the above-described transient control will be described with reference to FIG. The flow in FIG. 9 starts in the SI mode. In step S91 after the start, the PCM 10 determines various parameters (for example, water temperature, outside air temperature, engine load, engine speed, fuel injection timing, fuel pressure, ignition timing). The intake valve opening / closing timing and the exhaust valve opening / closing timing, etc.) are read to grasp the operating state of the engine 1. In step S92, it is determined whether or not the required load has decreased, that is, whether or not to shift from the SI mode to the CI mode. When the transition is not performed (that is, when NO), steps S91 and S92 are repeated, whereas when the transition is made to the CI mode (that is, when YES), the process proceeds to step S93. From the start of the flow to step S112 corresponds to the first cycle of the time charts shown in FIGS.

  In step S93, it is determined whether or not the load at the transfer destination is a medium load, that is, whether or not the transfer destination is the region (2). When the transfer destination is the area (2) (that is, when YES), the process proceeds to step S94, and when the transfer destination is the area (1) (that is, when NO), the process proceeds to step S99. The transition to step S94 corresponds to the second transient mode, and the transition to step S99 corresponds to the first transient mode.

  In step S94, in the SI mode, it is determined whether or not the intake amount reduction control is being performed by the throttling or the slow closing control of the intake valve 21, and when the reduction control has already been performed (that is, YES). ), The process proceeds to step S95. On the other hand, when the weight reduction control is not performed (that is, when NO), the process proceeds to step S96. In step S96, the intake air amount (fresh air amount) is reduced by the slow closing control of the intake valve 21 and / or the throttle control of the throttle valve 36. This reduces the effective compression ratio.

  In step S95, fuel injection is performed after the middle of the compression stroke, and compression ignition combustion is performed in subsequent step S97. Thus, as described above, the stability of the compression ignition combustion in the second transient mode is achieved. Steps S94 to S97 correspond to the second cycle of the time chart shown in FIG.

  Thus, when the second transient mode ends, the process proceeds to step S98, where the intake air amount reduction control is ended, the exhaust-side VVL 71 is turned on, and the exhaust double opening is started. Thus, the compression ignition combustion is performed while the effective compression ratio is relatively increased. This step S98 corresponds to the third cycle of the time chart shown in FIG.

  On the other hand, step S99 corresponds to the transition to the relatively low load side region (1), the throttle opening is fully opened, and the EGR valve 511 and the EGR cooler bypass valve 531 are closed in the subsequent step S910. Then, in step S911, fuel injection is performed in the intake stroke or the initial stage of the compression stroke, and as described above, the air-fuel mixture that is relatively homogeneous and has an A / F leaner than the stoichiometric air-fuel ratio is compressed and burned (step S912). Steps S99 to S912 correspond to the second cycle of the time chart shown in FIG.

  Thereafter, in step S913, the VVL 71 on the exhaust side is turned on, the exhaust is opened twice, and the A / F is made rich relative to steps S911 and S912 to perform compression ignition combustion. This step S913 corresponds to the third cycle of the time chart shown in FIG.

  Here, in the above configuration, when switching from the SI mode to the CI mode, the first transient mode and the second transient mode are interposed for only one cycle, but the exhaust gas temperature is reduced to a desired level. It may be determined whether or not it has decreased, and the first and second transient modes may be continued for a plurality of cycles until the level decreases to a desired level. The exhaust gas temperature may be estimated by the PCM 10 based on various parameters read, for example.

  Further, based on the operating state of the engine 1 before and after switching from the SI mode to the CI mode, the number of cycles for executing the first transient mode and the second transient mode is set in advance and stored in the PCM 10, When switching from the SI mode to the CI mode, the first transient mode and the second transient mode may be continued for the set number of cycles.

  In the above-described configuration, the valve mechanism of the intake valve 21 includes the VVL 73 that switches between the large lift cam and the small lift cam. However, the valve mechanism of the intake valve 21 is replaced with VVL. You may make it provide the lift amount variable mechanism (CVVL (Continuously VariableValve Lift)) which can change the lift amount continuously. CVVL can adopt various known structures as appropriate, and illustration of the detailed structure is omitted. With VVT and CVVL, the intake valve 21 can continuously change its valve opening timing and valve closing timing, and the lift amount (and valve opening period).

  FIGS. 10 and 11 respectively show switching control from the SI mode to the CI mode in a configuration in which the valve mechanism of the intake valve 21 includes CVVL. Of these, FIG. 10 relates to switching from the high load SI mode to the relatively high load region (2) in the CI mode, which corresponds to FIG. FIG. 11 relates to switching from the high load SI mode to the relatively low load region (1) in the CI mode, which corresponds to FIG.

  First, in the leftmost first cycle in FIG. 10, the operation is performed in the high load SI mode, and the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1). The CVVL of the intake valve 21 drives the intake valve 21 with a relatively small lift so that the fresh air amount is commensurate with the fuel injection amount, and the VVT 72 is relatively early before the intake bottom dead center. Set to the time. By thus closing the intake valve 21 early, the amount of fresh air is limited. Further, in the first cycle, the amount of fresh air that is not limited only by the control of the intake valve 21 is compensated by the throttle valve 36 being throttled. Further, the external EGR gas is introduced into the cylinder 18 by opening the EGR valve 511 and / or the EGR cooler bypass valve 531. The VVL 71 on the exhaust side is off and the internal EGR gas is not introduced.

  The second cycle corresponds to the second transient mode. In the second cycle, the throttle opening is reduced and the intake side VVT 72 sets the closing timing of the intake valve 21 to be delayed after the intake bottom dead center. As a result, the amount of fresh air introduced into the cylinder 18 is reduced. Further, the EGR valve 511 and / or the EGR cooler bypass valve 531 remain open, and external EGR gas is introduced into the cylinder 18. Further, the exhaust-side VVL 71 also remains off, and the internal EGR gas is not introduced into the cylinder 18. As a result, the effective compression ratio is lower than in the first cycle and lower than in the third cycle.

  Then, in the second cycle in which compression ignition combustion is performed, the compression end temperature and the compression end pressure are reduced by limiting the amount of fresh air and lowering the effective compression ratio. . On the other hand, compression ignition combustion is stably performed by performing fuel retard injection. As a result, the exhaust gas temperature exhausted after the second cycle combustion is lowered.

  The third cycle corresponds to the CI mode cycle after transition to the region (2). In the third cycle, the VVT 72 is operated to advance the closing timing of the intake valve 21 more than in the second cycle, and the exhaust-side VVL 71 is turned on to open the exhaust twice. Thereby, although a part of the burned gas generated by the compression ignition combustion in the second cycle is introduced into the cylinder 18, as described above, the temperature of the burned gas is suppressed to be low and 2 In the cycle, the temperature state in the cylinder is also kept low, so the temperature state in the cylinder 18 in the third cycle is not so high. Further, in the third cycle, while the throttle valve 36 that has been throttled in the second cycle is fully opened, the EGR valve 511 and / or the EGR cooler bypass valve 531 are continuously opened, so that the external EGR gas is released. It is introduced into the cylinder 18. In this way, the intake air (gas amount) introduced into the cylinder 18 in the third cycle increases compared to the second cycle, so the effective compression ratio in the third cycle is higher than the effective compression ratio in the second cycle. Become. Further, the EGR rate in the third cycle is higher than the EGR rate in the second cycle by the amount of internal EGR gas introduced.

  Thus, the compression ignition combustion is performed even in the third cycle. However, as described above, in the third cycle, the temperature state in the cylinder 18 does not become too high, and the ignition is performed prematurely by performing the retard injection. Without being compressed, the compression ignition is surely performed in the vicinity of the compression top dead center and the combustion is stably performed.

  Next, in the leftmost first cycle in FIG. 11, the operation is performed in the high load SI mode, and the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1). The CVVL of the intake valve 21 drives the intake valve 21 with a relatively small lift so that the fresh air amount is commensurate with the fuel injection amount, and the VVT 72 is relatively early before the intake bottom dead center. Limiting the amount of fresh air by setting the time. Further, in the first cycle, the throttle valve 36 is throttled, but the throttle valve 36 is gradually opened in the fully open direction as in FIG. In addition, the external EGR gas is introduced into the cylinder 18 and the VVL 71 of the exhaust valve 22 is off.

  In the second cycle, the throttle opening is set to fully open, while the VVT 72 on the intake side delays the closing timing of the intake valve 21 to the vicinity of the intake bottom dead center. As a result, the amount of fresh air introduced into the cylinder 18 increases. Also, the pumping loss is reduced. In the second cycle, the EGR valve 511 and the EGR cooler bypass valve 531 are fully closed to stop the introduction of the external EGR gas into the cylinder 18, but the relatively high-temperature exhaust gas remaining in the EGR passage 50 remains. Is introduced into the cylinder 18 in the second cycle. Thus, the effective compression ratio in the transient mode is relatively higher when shifting to the region (1) than when shifting to the region (2). In the second cycle, fuel is injected into the cylinder 18 during the period from the intake stroke to the initial compression stroke, and the air-fuel mixture that is relatively homogeneous and leaner than the stoichiometric air-fuel ratio self-ignites near the compression top dead center. And burn. As a result, the exhaust gas temperature decreases.

  In the third cycle, the VVL 71 on the exhaust side is turned on and the exhaust is opened twice, and the exhaust gas having a low temperature is introduced into the cylinder 18, so that the compression ignition combustion is not caused without premature ignition. Can be performed stably.

  Thus, also in the valve operating mechanism of the intake valve 21 including the CVVL, it is possible to perform the transient control for switching between the first transient mode and the second transient mode as described above. However, the VVL 73 is excellent in that the switching of the intake air can be instantaneously performed, the response of the transient control is improved, and the switching of the mode becomes smooth.

  The technique disclosed here is not limited to application to the engine configuration described above. For example, fuel may be injected into the intake port 16 through a port injector provided separately in the intake port 16 instead of the injector 67 provided in the cylinder 18 during the intake stroke period.

  The engine 1 is not limited to an in-line 4-cylinder engine, and may be applied to an in-line 3-cylinder, in-line 2-cylinder, in-line 6-cylinder engine, or the like. Further, the present invention can be applied to various engines such as a V type 6 cylinder, a V type 8 cylinder, and a horizontally opposed 4 cylinder.

  Moreover, the operation area | region shown in FIG. 4 is an illustration, and it is possible to provide various operation areas besides this.

1 Engine (Engine body)
10 PCM (controller)
18 cylinder 21 intake valve 22 exhaust valve 25 spark plug 36 throttle valve 67 injector (fuel injection valve)

Claims (7)

An engine body having a cylinder;
A fuel injection valve configured to inject fuel supplied into the cylinder;
A spark plug configured to ignite the air-fuel mixture in the cylinder;
A controller configured to operate the engine body by controlling at least the fuel injection valve and the spark plug; and
The controller is in a compression ignition mode for performing compression ignition combustion in which the air-fuel mixture is combusted by self-ignition when the engine body is in a predetermined operation region, and when in the region other than the predetermined operation region, In the spark ignition mode for performing spark ignition combustion for igniting and burning the air-fuel mixture by driving the spark plug,
The controller may also be configured to change the operation state of the engine body when the mode is switched when the operation state of the engine body shifts from the operation region in which the spark ignition mode is performed to the low load side region in the operation region in which the compression ignition mode is performed. When the mode is switched, when the transition mode of the spark ignition mode is interposed and the region is shifted from the operation region in which the spark ignition mode is performed to the region on the higher load side than the low load side region in the operation region in which the compression ignition mode is performed. Intervening the second transient mode,
The controller has an effective compression ratio at the time of execution of the second transient mode that is lower than an effective compression ratio set at a transition destination and is lower than an effective compression ratio at the time of execution of the first transient mode. Control device for spark-ignition engine set low. The control device for a spark ignition engine according to claim 1,
The controller is a controller for a spark ignition engine that reduces the effective compression ratio by reducing intake air into the cylinder when the second transient mode is executed. In the control device for the spark ignition engine according to claim 1 or 2,
The controller is a control device for a spark ignition engine that sets the air-fuel ratio of the air-fuel mixture to lean when executing the first transient mode as compared to when executing the second transient mode. In the control device for the spark ignition engine according to any one of claims 1 to 3,
The controller is a control device for a spark ignition engine that sets an EGR rate at the time of execution of the second transient mode to be lower than an EGR rate set at a transition destination. The control device for the spark ignition engine according to any one of claims 1 to 4,
When the controller restricts intake air in the spark ignition mode before execution of the second transient mode, the controller for the spark ignition engine that continues the intake air restriction even in the second transient mode . The control device for the spark ignition engine according to claim 5,
The controller is a control device for a spark ignition engine that continues the control even in the second transition mode when the throttle ring or the intake valve is slowly closed in the spark ignition mode. In the control device for a spark ignition engine according to any one of claims 1 to 6,
The controller is a control device for a spark ignition engine that sets the fuel injection timing of the fuel injection valve after the middle of the compression stroke when the second transient mode is executed.

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