JP5589956B2 - Compression ignition gasoline engine - Google Patents

Description

  The technology disclosed herein relates to a compression ignition gasoline engine.

  A compression ignition type gasoline engine that compresses and ignites a lean air-fuel mixture is known as a technology that achieves both improved exhaust emission and improved thermal efficiency. For example, Patent Document 1 discloses that in this compression ignition type gasoline engine, in order to stabilize the compression ignition by increasing the in-cylinder temperature, the exhaust valve is restarted during the intake stroke so that a relatively high temperature burned gas is generated. A technique to be introduced into the cylinder is described.

  On the other hand, as the engine load increases, the compression ignition combustion becomes premature ignition combustion in which the pressure rises rapidly. As a result, combustion noise increases and abnormal combustion such as knocking occurs, and Raw NOx increases due to a high combustion temperature. Thus, for example, Patent Literature 2 describes a gasoline engine that performs compression ignition combustion only in a low load side operation region and performs spark ignition combustion by driving a spark plug in a high load side operation region. . Further, for example, in Patent Document 3, when switching from compression ignition combustion to spark ignition combustion, knocking is avoided by temporarily setting the air-fuel ratio to be rich because the cylinder temperature is too high and knocking is likely to occur. The technology to do is described.

JP 2007-132319 A JP 2007-154859 A JP 2009-91994 A

  As described above, in compression ignition combustion, since stable combustion is required particularly in a low load region, it is preferable to increase the in-cylinder temperature. On the other hand, when switching between compression ignition combustion and spark ignition combustion, the cylinder temperature that is too high during spark ignition combustion causes knocking (that is, pre-ignition).

  The technology disclosed herein has been made in view of the above points, and an object thereof is to stabilize compression ignition combustion in a compression ignition gasoline engine that performs compression ignition combustion.

  Specifically, the compression ignition type gasoline engine disclosed herein controls at least the opening operation of the engine body configured to supply fuel containing at least gasoline into the cylinder, and the intake valve and the exhaust valve. And the controller opens the exhaust valve during the intake stroke when the engine body is at least warmed up and its operating state is in a predetermined low load range. Thus, while introducing the burned gas into the cylinder, the compression ignition mode in which the air-fuel mixture in the cylinder is compressed and ignited, and when the operating state of the engine body is in a higher load region than the compression ignition mode, The engine body is operated with the valve opening operation of the exhaust valve stopped during the intake stroke so that introduction of burned gas into the cylinder is substantially stopped.

  Then, among the intake port and the exhaust port communicating with the cylinder, at least the port where the exhaust valve that opens during the intake stroke in the compression ignition mode is arranged is directed toward the inside of the cylinder. A heating means for heating the gas passing through the port is provided.

  Here, “a port where an exhaust valve that opens during the intake stroke in the compression ignition mode” is provided when a plurality of exhaust valves are provided for one cylinder, and during the intake stroke. When the exhaust valve to be opened is one of them, as a specific example, there are two exhaust valves, and when only one of the exhaust valves opens during the intake stroke, the one exhaust valve Is the port where is located. That is, not all exhaust ports are provided with heating means.

  “To stop the opening of the exhaust valve during the intake stroke so that the introduction of burned gas into the cylinder is substantially stopped” means that the exhaust valve is closed during the intake stroke. In addition to this, if the introduction of burned gas into the cylinder is substantially stopped, this includes that the exhaust valve is slightly opened during the intake stroke.

  First, when the operating state of the engine body is in a predetermined low load region, the compression ignition mode is executed in which compression ignition combustion is executed. Compression ignition combustion is advantageous for improving both exhaust emission and thermal efficiency. In the compression ignition mode, the engine body may be operated with leaner than the stoichiometric air-fuel ratio. In the compression ignition mode, a relatively high temperature burned gas is introduced into the cylinder by opening the exhaust valve during the intake stroke. That is, the compression ignition combustion is stabilized by increasing the compression end temperature by executing the internal EGR control. In addition, when the engine operating region is in a region where the load is higher than in the compression ignition mode, the opening operation of the exhaust valve during the intake stroke is stopped, thereby substantially stopping the introduction of burned gas. To do. By doing so, it is avoided that the in-cylinder temperature becomes too high in a relatively high load region, which is advantageous in avoiding knocking.

  Thus, the exhaust port or the intake and exhaust ports are provided with heating means for heating the gas passing through the ports toward the inside of the cylinder. The heating means provided in the exhaust port heats the burned gas introduced into the cylinder and raises the temperature when the exhaust valve opens during the intake stroke. On the other hand, when the heating means is provided in the intake port, the air introduced into the cylinder during the intake stroke is heated to raise the temperature. As a result, in the compression ignition mode, the in-cylinder temperature is increased and the compression ignition combustion is stabilized.

  Here, as described above, the heating means provided in the exhaust port heats the burned gas introduced into the cylinder when the exhaust valve is opened during the intake stroke and raises the temperature. In an operation region where the load is high, the opening operation of the exhaust valve during the intake stroke is stopped and the introduction of burned gas is substantially stopped, so that the in-cylinder temperature is not increased. This is advantageous in avoiding that the temperature in the cylinder becomes too high in the operation region where the load is relatively high.

  On the other hand, when the heating means is provided in the intake port, it is desirable that the temperature of the air introduced into the cylinder is not increased by the heating means in the operation region where the load is relatively high.

  Therefore, the heating means may be provided only in the exhaust port among the intake port and the exhaust port.

  In this way, since no heating means is provided at the intake port, the temperature of the air introduced into the cylinder (that is, fresh air) is not increased in the operation region where the load is relatively high. On the other hand, the heating means provided at the exhaust port stops the opening operation of the exhaust valve during the intake stroke in an operation region where the load is relatively high, so the in-cylinder temperature does not increase.

  In particular, when the heating means is provided in the intake port, in the compression ignition mode, while the air is heated by the heating means, immediately after shifting from the compression ignition mode to the operation region with a relatively high load, to the heating means, For example, even if energization is stopped and heating is stopped, the temperature of the air introduced into the cylinder may increase due to the residual heat of the heating means. On the other hand, in the configuration in which the heating means is not provided in the intake port, the temperature rise of the introduced air due to residual heat is reliably avoided. Therefore, the configuration in which the heating means is provided only at the exhaust port is advantageous in making it more reliable to avoid knocking in the operation region where the load is relatively high.

In the above configuration, the heating means includes a heating element that is switched between operation and non-operation by the controller, and the controller is configured such that when the operating state of the engine body is in the high load region, It stops the operation of the heating element.

  When the heating means includes the heating element, the gas introduced into the cylinder can be efficiently heated, and the compression end temperature is more reliably increased. This is advantageous for stabilization of compression ignition combustion. On the other hand, when the operating state of the engine body is in a relatively high load region, the operation of the heating element is stopped. This avoids an excessive increase in the cylinder temperature. Further, in the configuration in which the heating means is provided in the exhaust port, as described above, the exhaust valve is stopped during the intake stroke, and the burned gas is prohibited from being introduced into the cylinder. Even if the body is operated, the temperature in the cylinder does not rise, but stopping the operation of the heating element is advantageous in reducing energy loss and, in turn, in improving fuel efficiency.

  The controller sets a spark ignition mode in which an air-fuel mixture in the cylinder is ignited by driving an ignition plug when the operating state of the engine body is at least in a fully open load region, and the ignition timing is burned in an expansion stroke. It is also possible to retard the angle so as to start. Thereby, a high torque is ensured while avoiding knocking at least in the fully open load region.

  The controller is configured so that the introduction of burned gas into the cylinder is substantially stopped when the operating state of the engine body is in a medium load region adjacent to the load region in the compression ignition mode. The air-fuel mixture in the cylinder may be compressed and ignited with the valve opening operation of the exhaust valve being stopped during the intake stroke.

  As a result, as the engine load increases, the compression ignition combustion becomes the combustion of premature ignition with a severe pressure rise, but in the middle load region of the engine, the exhaust valve opening operation during the intake stroke is performed. By suppressing the increase in the in-cylinder temperature by stopping, compression ignition combustion becomes possible while avoiding premature ignition. Therefore, the compression ignition mode region is expanded correspondingly to the high load side, which is advantageous for improving exhaust emission and improving thermal efficiency. In this load region, external EGR control is performed to introduce the relatively low-temperature burned gas into the cylinder by recirculating the burned gas through the EGR passage connecting the exhaust side and the intake side. Also good.

  The heating means may include a heat storage material. In this way, the heating means provided in the exhaust port in particular allows the heat storage material to absorb heat when the burned gas is exhausted from the cylinder while the heat storage material stores heat when the burned gas is introduced into the cylinder. The material dissipates heat and raises the temperature of the burned gas. In other words, although the temperature of the burned gas once discharged from the cylinder to the intake port is slightly lowered until it is reintroduced into the cylinder, Heat can be used effectively. This reduces the additional energy required to increase the in-cylinder temperature as much as possible or eliminates the need for additional energy.

  The heating means may be provided at a throat of the port. Since the throat is the place where the gas flow rate is the highest, disposing the heating means here can provide a high heat transfer coefficient and efficiently raise the temperature of the gas. In addition, providing the heating means including the above-described heat storage material at the throat of the exhaust port increases the heat absorption efficiency and the heat dissipation efficiency for the burned gas that passes through the throat. Make it even less. Further, since the heat source is not exposed in the combustion chamber, it is advantageous for avoiding knocking.

  As described above, this compression ignition type gasoline engine is provided with a heating means at least in a port where an exhaust valve that opens during the intake stroke in the compression ignition mode is provided. Since the temperature of the introduced gas is increased by heating the cylinder, it is possible to increase the temperature in the cylinder, so that the compression ignition combustion is stabilized, while in the region where the load is higher than the compression ignition mode, the intake stroke By stopping the opening operation of the exhaust valve inside and substantially stopping the introduction of the burned gas into the cylinder, it is advantageous for avoiding knocking by avoiding that the temperature in the cylinder becomes too high. .

It is the schematic which shows the structure of a spark ignition type gasoline engine. It is a block diagram concerning control of a spark ignition type gasoline engine. It is a figure which illustrates the operating area of an engine. It is a figure which compares the state of SI combustion by a high pressure retarded injection, and the state of conventional SI combustion. It is a timing chart which shows the difference in operation | movement of an intake valve and an exhaust valve with respect to the difference in an engine load, and the difference in an ignition timing and an injection timing. (A) Mixture filling amount, (b) Throttle valve opening, (c) EGR when the internal EGR amount is controlled by controlling the intake valve in the low load region and the external EGR is used in the high load region Change in valve opening, (d) closing timing of exhaust valve opening twice, (e) opening timing of intake valve, (f) closing timing of intake valve, (g) lift amount of intake valve It is a figure which shows an example, respectively. (A) Mixture filling amount, (b) G / F, (c) External EGR rate, (d) when controlling the internal EGR amount in the low load range and using the external EGR in the high load range It is a figure which shows an example of the change of injection timing, (e) fuel pressure, (f) injection pulse width, and (g) ignition timing, respectively. FIG. 10 is an enlarged cross-sectional view (sectional view taken along line VIII-VIII in FIG. 9) showing the vicinity of an intake port and an exhaust port of the engine. It is plane explanatory drawing which shows the structure of the heater provided in the exhaust port. It is a figure explaining an example of the lift curve of an intake valve and an exhaust valve, and the flow of burned gas at that time. FIG. 10 is a diagram corresponding to FIG. 9 illustrating a heater having a configuration different from that of FIG. 9.

  Hereinafter, an embodiment of a control device for a spark ignition gasoline 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 is provided with a cylinder block 11 provided with a plurality of cylinders 18 (only one shown), a cylinder head 12 provided on the cylinder block 11, and a cylinder block 11 below the cylinder block 11. And an oil pan 13 in which oil is stored. 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 having a reentrant shape is formed on the top surface of the piston 14. The cavity 141 faces a direct injection 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 14 or more for the purpose of improving the theoretical thermal efficiency, stabilizing the compression ignition combustion described later, and the like. In addition, what is necessary is just to set a geometric compression ratio suitably in the range of about 14-20.

  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 VVL 71 is not shown in detail in its configuration, two types of cams having different cam profiles, a first cam having one cam crest and a second cam having two cam crests, and its first And a lost motion mechanism that selectively transmits an operating state of one of the 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 a normal mode in which the valve is opened only once during the exhaust stroke (see FIGS. 5B, 5C, and 5D). On the other hand, when the operating state of the second cam is transmitted to the exhaust valve 22, the exhaust valve 22 opens during the exhaust stroke and also opens during the intake stroke. It operates in a special mode that opens twice (see FIG. 5A). The normal mode and the special mode of the VVL 71 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 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. Also, the execution of internal EGR is not realized only by opening the exhaust twice. For example, the internal EGR control may be performed by opening the intake valve 21 twice or by opening the intake valve twice, or by providing 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. Internal EGR control that causes the fuel gas to remain in the cylinder 18 may be performed.

  As shown in FIG. 2, the exhaust side valve system having the VVL 71 is arranged on the intake side as shown in FIG. 2. 72) and a lift variable mechanism (hereinafter referred to as CVVL (Continuously Variable Valve Lift)) 73 capable of continuously changing the lift amount of the intake valve 21. . The VVT 72 may employ a hydraulic, electromagnetic, or mechanical structure as appropriate, and illustration of the detailed structure is omitted. The CVVL 73 can also adopt various known structures as appropriate, and the detailed structure is not shown. With the VVT 72 and the CVVL 73, the intake valve 21 can change its valve opening timing, valve closing timing, and lift amount, as shown in FIGS.

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

  As shown in an enlarged view in FIG. 8, the direct injection injector 67 is disposed such that its injection hole faces the inside of the combustion chamber 19 from the central portion of the ceiling surface of the combustion chamber 19. The direct injection injector 67 directly injects an amount of fuel into the combustion chamber 19 at an injection timing according to the operating state of the engine 1 and according to the operating state of the engine 1. In this example, the direct injection injector 67 is a multi-injector type injector having a plurality of injection holes, although detailed illustration is omitted. Thereby, the direct injection injector 67 injects the fuel so that the fuel spray spreads radially. At the timing when the piston 14 is positioned near the compression top dead center, the fuel spray injected radially from the central portion of the combustion chamber 19 flows along the wall surface of the cavity 141 formed on the top surface of the piston. As a result, it reaches around the spark plug 25 described later. 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. The combination of the multi-hole injector 67 and the cavity 141 is advantageous in that it shortens the time until the fuel spray reaches around the spark plug 25 after fuel injection and shortens the combustion period. is there. In addition, the direct injection injector 67 is not limited to a multi-injection type injector, and an external valve-opening type injector may be adopted as the direct injection injector.

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

  The fuel tank (not shown) and the direct injection injector 67 are connected to each other by a high-pressure fuel supply path. A high-pressure fuel supply system 62 that includes a high-pressure fuel pump 63 and a common rail 64 and supplies fuel to the direct injection injector 67 at a relatively high fuel pressure is interposed on the high-pressure fuel supply path. The high-pressure fuel pump 63 pumps fuel from the fuel tank to the common rail 64, and the common rail 64 stores the pumped fuel at a high fuel pressure. When the direct injection injector 67 is opened, the fuel stored in the common rail 64 is injected from the injection port of the direct injection injector 67. Here, although not shown, the high-pressure fuel pump 63 is a plunger type pump, and is driven by the engine 1 by being connected to a timing belt between a crankshaft and a camshaft, for example. The high-pressure fuel supply system 62 configured to include this engine-driven pump makes it possible to supply fuel with a high fuel pressure of 40 MPa or more to the direct injection injector 67. The pressure of the fuel supplied to the direct injection injector 67 is changed according to the operating state of the engine 1 as will be described later. The high-pressure fuel supply system 62 is not limited to this.

  Similarly, the fuel tank (not shown) and the port injector 68 are connected to each other by a low-pressure fuel supply path. On this low pressure fuel supply path, a low pressure fuel supply system 66 for supplying fuel with a relatively low fuel pressure to the port injector 68 is interposed. Although not shown in detail, the low-pressure fuel supply system 66 includes an electric or engine-driven low-pressure fuel pump and a regulator, and is configured to supply a predetermined pressure of fuel to each port injector 68. . Since the port injector 68 injects fuel into the intake port, the pressure of the fuel supplied by the low pressure fuel supply system 66 is set lower than the pressure of the fuel supplied by the high pressure fuel supply system 62.

  A spark plug 25 that ignites the air-fuel mixture in the combustion chamber 19 is also attached to the cylinder head 12. The spark plug 25 is disposed through the cylinder head 12 so as to extend obliquely downward from the exhaust side of the engine 1. As shown in FIG. 8, the tip of the spark plug 25 is disposed facing the combustion chamber 19 in the vicinity of the tip of the direct injection injector 67 disposed in the center portion of the combustion chamber 19.

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

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

  Between the air cleaner 31 and the surge tank 33 in the intake passage 30, a water-cooled intercooler / warmer 34 that cools or heats the air and a throttle valve 36 that adjusts the amount of intake air to each cylinder 18 are arranged. 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. The temperature of fresh air introduced into the cylinder 18 is adjusted by adjusting the ratio between the flow rate of the intercooler bypass passage 35 and the flow rate of the intercooler / warmer 34 through the adjustment of the opening degree of the intercooler bypass valve 351.

  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 and SW15, and a common rail 64 of the high pressure fuel supply system 62 are attached. Further, a fuel pressure sensor SW for detecting the fuel pressure supplied to the direct injection injector 67 16.

  The PCM 10 performs various calculations based on these detection signals to determine the state of the engine 1 and the vehicle, and in response to this, the direct injection injector 67, the port injector 68, the spark plug 25, the VVT 72 on the intake valve side, Control signals are output to the actuators of the CVVL 73, the exhaust valve side VVL 71, the high-pressure fuel supply system 62, and 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. 3 shows an example of the operation region of the engine 1. For the purpose of improving fuel efficiency and exhaust emission, the engine 1 is compressed and self-ignited without ignition by the spark plug 25 in a low load region where the engine load is relatively low after the engine 1 is warmed up. Compressed ignition combustion is performed by performing combustion. 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. Therefore, the engine 1 is configured to switch between 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 according to the operating state of the engine 1.

  As will be described in detail later, in the CI mode, for example, a relatively homogeneous lean mixing is achieved by injecting fuel into the cylinder 18 by the direct injection injector 67 at a relatively early timing, for example, during an intake stroke or a compression stroke. The air-fuel mixture is compressed and self-ignited in the vicinity of the compression top dead center. In the relatively low load region in the CI mode, the internal EGR control is used together from the viewpoint of stabilizing the compression ignition combustion by increasing the in-cylinder temperature. On the other hand, in the relatively high load region in the CI mode, as the engine load increases, the in-cylinder temperature rises and the compression ignition combustion becomes steep, so the internal EGR control is stopped and switched to the external EGR control. In the middle load region, the internal EGR control and the external EGR control may be used together in part.

  In contrast, in the SI mode, basically, for example, during the intake stroke or the compression stroke, the direct injection injector 67 injects fuel into the cylinder 18 to form a homogeneous or stratified air-fuel mixture and improve the compression. The mixture is ignited by performing ignition in the vicinity of the dead center. In the SI mode, the engine 1 is also operated at the theoretical air fuel ratio (λ = 1). This makes it possible to use a three-way catalyst, which is advantageous for improving the emission performance.

  The geometric compression ratio of the engine 1 is set to 14 or more (for example, 18) as described above. Since the high compression ratio increases the compression end temperature and the compression end pressure, the CI mode is advantageous in stabilizing the compression ignition combustion. On the other hand, since the high compression ratio engine 1 is switched to the SI mode in the high load region, the higher the engine load, the faster the ignition or the knocking (hereinafter, these may be collectively referred to as knocking). There is an inconvenience that abnormal combustion tends to occur.

  Therefore, in this engine 1, when the operating state of the engine is in a high load range, abnormal combustion is avoided by executing SI combustion in which the fuel injection form is greatly different from the conventional one. Specifically, this fuel injection mode is a period that is significantly retarded from the latter half of the compression stroke to the early stage of the expansion stroke with a fuel pressure that is significantly higher than in the past (hereinafter, this period is referred to as the retard period). The fuel is injected into the cylinder 18 by the direct injection injector 67. This characteristic fuel injection mode is hereinafter referred to as “high pressure retarded injection”.

  FIG. 4 shows the heat generation rate (upper figure) and the progress of the unburned mixture reaction in the SI combustion by the high pressure retarded injection (solid line) 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. 4 is the crank angle. As a premise for this comparison, the operating state of the engine 1 is a high load region in the low speed region, and the amount of fuel to be injected is the same in the case of SI combustion by high pressure retarded injection and conventional SI combustion.

  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. 4, the combustion ends through the peak of the heat generation rate. Here, the period 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 the reaction possible time). As shown by the broken line in the lower diagram of FIG. During this time, the reaction of the unburned mixture gradually proceeds. The dotted line in the figure shows the ignition threshold, which is the degree of reactivity with which the unburned mixture reaches ignition. Conventional SI combustion has a very long reaction time, during which the unburned mixture reacts. Since it continues to advance, the reactivity of the unburned mixture before and after ignition exceeds the ignition threshold, causing abnormal combustion such as premature ignition or knocking.

  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. 4, the reaction possible time is the period during which the direct injection injector 67 injects the fuel ((1) injection period) and the combustible air-fuel mixture around the spark plug 25 after the completion of the injection. The sum of the period until formation ((2) mixture formation period) and the period until combustion ended by ignition ((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.

  That is, a higher turbulence energy in the cylinder 18 is advantageous for shortening the combustion period. As described above, fuel injection at a high fuel pressure increases the turbulent energy in the cylinder 18, but if the injection timing is during the intake stroke, the time until the ignition timing is long, and the compression after the intake stroke The disturbance is attenuated due to the compression of the cylinder 18 in the stroke, and the disturbance energy in the cylinder during the combustion period becomes relatively low. That is, even if fuel is injected into the cylinder 18 at a high fuel pressure, the combustion period is hardly shortened as long as the injection timing is during the intake stroke.

  On the other hand, injecting fuel into the cylinder at a timing within the retard period, that is, at a relatively late timing and at high fuel pressure, starts combustion while suppressing attenuation of turbulence in the cylinder. to enable. For this reason, the turbulent energy in the cylinder during the combustion period increases. This shortens the combustion period and stabilizes combustion.

  Moreover, a high fuel pressure relatively increases the amount of fuel injected per unit time. For this reason, when compared with the same fuel injection amount, a high fuel pressure shortens the injection period compared with a low fuel pressure.

  Furthermore, the high fuel pressure is advantageous for atomizing the fuel spray injected into the cylinder, and makes the flight distance of the fuel spray longer. Thus, high fuel pressure shortens the mixture formation period.

  Therefore, the shortening of the injection period and the mixture formation period described above make it possible to set the fuel injection timing (more precisely, the injection start timing) to a relatively late timing. That is, the high fuel pressure enables fuel injection within the retard period.

  Thus, in the spark ignition mode, by performing fuel injection into the cylinder at a high fuel pressure and within the retard period, rapid combustion with a short combustion period is realized, thereby stabilizing the combustion.

  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. 4, 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 the reaction possible time, in the conventional intake stroke injection at a low fuel pressure, the reaction progress of the unburned mixture at the end of combustion exceeds the ignition threshold, and abnormal combustion occurs. As a result, the high-pressure retarded injection suppresses the progress of the reaction of the unburned air-fuel mixture at the end of combustion, and can avoid abnormal combustion.

  Next, with reference to FIG. 5, a description will be given of an operation state of the intake valve 21 and the exhaust valve 22 and a control example of the fuel injection timing and the ignition timing corresponding to the operating state of the engine 1. Here, in FIGS. 5A, 5B, 5C, and 5D, the engine load increases in the order of (a) <(b) <(c) <(d). (A) and (b) are low and medium load regions corresponding to the CI mode, and (c) is a high load region corresponding to the SI mode. (D) is a fully open load region corresponding to the SI mode.

  First, Fig.5 (a) shows the time of the driving | running state of the engine 1 in a low load region. Since this operation region is in the CI mode, the exhaust valve 22 is opened twice during the intake stroke under the control of the VVL 71 (see the solid line Ex2 in the figure. The solid line is the exhaust valve 22). ), And the broken line indicates the lift curve of the intake valve 21), whereby the internal EGR gas is introduced into the cylinder 18. The introduction of internal EGR gas increases the compression end temperature and stabilizes compression ignition combustion. However, since the temperature in the cylinder 18 naturally increases as the engine load increases, the internal EGR amount is reduced from the viewpoint of avoiding premature ignition (that is, knocking). As illustrated in FIG. 5, the internal EGR amount may be adjusted by adjusting the lift amount of the intake valve 21 under the control of the CVVL 73. Although not shown in FIG. 5, the internal EGR amount may be adjusted by adjusting the opening degree of the throttle valve 36. The fuel injection timing is set during the intake stroke, and the direct injection injector 67 injects fuel into the cylinder 18 to form a homogeneous lean air-fuel mixture in the cylinder 18. The fuel injection amount is set according to the load of the engine 1.

  FIG. 5A shows an example in which the valve opening period of the exhaust valve 22 during the intake stroke is set in the first half of the intake stroke. The valve opening period of the exhaust valve 22 may be set in the latter half of the intake stroke. When the valve opening period is set to the first half of the intake stroke, the exhaust valve 22 may be left open from the exhaust stroke sandwiching the exhaust top dead center to the first half of the intake stroke.

  FIG. 5B shows a case where the operating state of the engine 1 is in an intermediate load region where the engine load is higher than that in FIG. Although this operation region is also in the CI mode, the engine load is high and the temperature in the cylinder is relatively high. Therefore, if the internal EGR gas is introduced into the cylinder 18, the temperature in the cylinder becomes too high. Therefore, when the engine load is in the middle load range, the exhaust valve 22 is stopped twice to stop the introduction of the internal EGR gas, and the external EGR gas is introduced into the cylinder 18 instead. The fuel injection timing is set to an appropriate timing during the intake stroke or the compression stroke. At this timing, the direct injection injector 67 injects fuel into the cylinder 18 to form a homogeneous or stratified lean mixture in the cylinder 18. The point that the fuel injection amount is set according to the load of the engine 1 is the same as in FIG.

  FIG. 5C shows a case where the operating state of the engine 1 is in a high load range. This operation region is the SI mode, and in this operation region, the opening of the exhaust valve 22 is stopped twice. In the SI mode, the filling amount is adjusted so that the air-fuel ratio λ = 1. The adjustment of the filling amount may be performed by slow closing of the intake valve 21 in which the throttle valve 36 is fully opened while the closing timing of the intake valve 21 is set after the intake bottom dead center by the control of the VVT 72 and the CVVL 73. Good. This is advantageous for reducing pump loss. The adjustment of the filling amount may also be performed by adjusting the opening amount of the EGR valve 511 and adjusting the amount of fresh air introduced into the cylinder 18 and the amount of external EGR gas while the throttle valve 36 is fully opened. . This is effective for reducing pumping loss and cooling loss. In addition, introduction of external EGR gas contributes to avoiding abnormal combustion and also has an advantage of suppressing generation of Raw NOx. Further, as the adjustment of the filling amount, the slow closing control of the intake valve 21 and the control of the external EGR may be combined. In particular, on the low load side in the high load region, in order to prevent the EGR rate from being too high, the charging amount is adjusted by the slow closing control of the intake valve 21 while introducing the external EGR into the cylinder 18. Also good.

  The form of fuel injection is the high-pressure retarded injection described above. Therefore, the direct injection injector 67 directly injects the fuel into the cylinder 18 with a high fuel pressure during the retard period from the latter half of the compression stroke to the early stage of the expansion stroke. The high-pressure retarded injection may be configured by one injection (that is, batch injection), but as shown in FIG. 5C, two injections of the first injection and the subsequent second injection are performed. It may be configured to perform within the retard period (that is, divided injection). Since the first injection can ensure a relatively long air-fuel mixture formation period, it is advantageous for fuel vaporization atomization. Since a sufficient mixture formation period is secured by the first injection, the injection timing of the second injection can be set to a more retarded timing. This is advantageous for improving the turbulent energy in the cylinder and for shortening the combustion period. When performing split injection, it is preferable to set the fuel injection amount of the second injection to be larger than the fuel injection amount of the first injection. By doing so, the turbulent energy in the cylinder 18 is sufficiently increased, which is advantageous for shortening the combustion period and for avoiding abnormal combustion. Such split injection is performed only on the relatively high load side where the fuel injection amount increases in the high load region, and batch injection is performed on the low load side in the high load region where the fuel injection amount is relatively small. It may be. Further, the number of divisions is not limited to two, and may be set to three or more.

  Thus, in the SI mode, ignition by the spark plug 25 is executed in the vicinity of the compression top dead center after the end of fuel injection.

  FIG. 5D shows a case where the operating state of the engine 1 is in the fully open load region. This operating region is the SI mode, as in FIG. 5C, and stops the exhaust valve 22 from being opened twice. Further, since it is in the fully open load region, the external EGR is also stopped by closing the EGR valve 511.

  The form of fuel injection is basically high-pressure retarded injection, and is constituted by two injections into the cylinder 18 during the retard period of the first injection and the second injection, as shown in the figure. . The high pressure retarded injection may be batch injection. In this fully open load region, injection during the intake stroke may be added for the purpose of improving the intake charge efficiency. In this intake stroke injection, the intake air charging efficiency is improved by the intake air cooling effect accompanying the fuel injection, which is advantageous in improving the torque. Therefore, when the operating state of the engine 1 is in the fully open load region, three fuel injections of the intake stroke injection and the first and second injections are executed, or the intake stroke injection and the batch injection are performed. Two fuel injections are performed.

  Here, as described above, the high pressure retarded injection in which the fuel is directly injected into the cylinder 18 by the direct injection injector 67 has a very high fuel pressure. Therefore, if fuel is directly injected into the cylinder 18 during the intake stroke with such a high fuel pressure, a large amount of fuel may adhere to the wall surface in the cylinder 18 and cause problems such as oil dilution. There is. Therefore, in this intake stroke injection, fuel is injected into the intake port 16 not through the direct injection injector 67 but through the port injector 68 that injects fuel with a relatively low fuel pressure. By doing so, the above-mentioned problems such as oil dilution can be avoided.

  6 shows the parameters of the engine 1 with respect to the load variation, that is, (b) the opening degree of the throttle valve 36, (c) the opening degree of the EGR valve 511, and (d) the closing timing of the exhaust valve 22 being opened twice. (E) The valve opening timing of the intake valve 21, (f) the valve closing timing of the intake valve 21, and (g) the lift amount of the intake valve are shown.

  FIG. 6A shows the state in the cylinder 18. This figure shows the configuration of the air-fuel mixture in the cylinder with the horizontal axis representing torque (in other words, engine load) and the vertical axis representing the air-fuel mixture charge amount in the cylinder. As described above, the area on the left side of the figure where the load is relatively low is the CI mode, and the area on the right side of the figure where the load is higher than the predetermined load is the SI mode. The fuel amount (total fuel amount) is increased as the load increases regardless of the CI mode and the SI mode. A fresh air amount for setting the stoichiometric air-fuel ratio (λ = 1) is set with respect to this fuel amount, and this new air amount increases with an increase in the fuel amount as the load increases. It will be.

  In the CI mode, as described above, since the internal EGR gas is introduced into the cylinder 18 on the low load side, the remaining amount of the filling amount is constituted by the internal EGR gas and excess fresh air. Further, on the high load side (that is, medium load) in the CI mode, the introduction of the internal EGR gas is stopped, while the external EGR gas is introduced into the cylinder 18, and the remaining filling amount is the same as the external EGR gas. Composed of surplus fresh air. Therefore, in the CI mode, a lean mixture is obtained. On the other hand, in the SI mode, the engine 1 is operated so that λ = 1, and external EGR gas is introduced into the cylinder 18.

  The throttle valve 36 is basically set to fully open regardless of the load of the engine 1 as shown in FIG. Further, as shown in FIG. 5C, the EGR valve 511 is set to be fully closed while the internal EGR gas is being introduced, while being fully opened when the external EGR gas is being introduced in the CI mode. Is done. Further, in the SI mode with a higher load than that, the EGR valve 511 is gradually closed as the load of the engine 1 increases, and is fully closed at the fully opened load. By adjusting the opening degree of the throttle valve 36 and the EGR valve 511, the external EGR rate becomes 0 (zero) on the low load side of the CI mode as shown in FIG. In the SI mode, as the load on the engine 1 increases, the EGR rate gradually decreases and becomes 0 again at the fully opened load.

  FIG. 6D shows the valve closing timing when the exhaust valve 22 is opened twice. When the internal EGR gas is introduced into the cylinder 18 in the CI mode, the valve closing timing is set to a predetermined timing between the exhaust top dead center and the intake bottom dead center. On the other hand, in the CI mode and the SI mode, when the internal EGR gas is not introduced, the valve closing timing is set to the exhaust top dead center.

  As described above, as shown in FIG. 6 (e), when the exhaust valve 22 is opened twice, which is set to a predetermined valve closing timing for the execution of the internal EGR, the valve opening timing of the intake valve 21 is As the load on the engine 1 increases, the engine 1 is advanced so as to approach the exhaust top dead center. Therefore, the internal EGR gas introduced into the cylinder 18 increases as the load on the engine 1 decreases, whereas the internal EGR gas introduced into the cylinder 18 decreases as the load on the engine 1 increases. . As the load on the engine 1 is lower, the compression end temperature in the cylinder 18 is increased by a large amount of internal EGR gas, which is advantageous in realizing stable compression ignition combustion. On the other hand, the higher the load of the engine 1, the lower the compression end temperature in the cylinder 18 by suppressing the internal EGR gas, which is advantageous in suppressing premature ignition. As the engine load increases and the internal EGR control is stopped, the opening timing of the intake valve 21 is set to the exhaust top dead center. In addition, as shown in FIG. 6F, the closing timing of the intake valve 21 is made constant at the intake bottom dead center.

  Furthermore, as shown in FIG. 6 (g), the lift amount of the intake valve 21 gradually increases from the minimum lift amount as the engine load increases during the control of the internal EGR, whereas the control of the internal EGR is performed. When not performed, the maximum lift amount is set constant regardless of the engine load.

  Therefore, in the SI mode, the throttle valve 36 is made fully constant (FIG. 6B), and the opening timing and closing timing of the intake valve 21 are made constant (FIGS. 6E and 6F). The lift amount is made constant at the maximum (FIG. 6 (g)). From this, the ratio between the amount of fresh air introduced into the cylinder 18 and the amount of external EGR gas is adjusted by adjusting the opening of the EGR valve 511. Such control is advantageous in reducing pump loss.

  In FIG. 6, the throttle valve 36 is fully opened on the low load side of the CI mode, while the opening timing of the intake valve 21 is advanced so as to approach the exhaust top dead center as the engine load increases. In contrast, the lift amount of the intake valve 21 is gradually increased. On the other hand, on the low load side of the CI mode, the throttle valve 36 is controlled so that the opening degree of the throttle valve 36 gradually increases as the engine load increases. The opening timing of the intake valve 21 may be constant at the exhaust top dead center and the lift amount of the intake valve 21 may be constant at a predetermined lift amount regardless of the engine load. By doing so, when the internal EGR control is being executed, the internal EGR gas amount introduced into the cylinder 18 is adjusted according to the opening degree of the throttle valve 36.

  Next, FIG. 7 shows each control parameter of the engine 1 with respect to the load fluctuation in the low speed range, that is, (b) G / F, (c) external EGR rate, (d) injection timing, (e) fuel pressure, ( f) Fuel injection pulse width (that is, injection period) and (g) changes in ignition timing.

  First, the state of the air-fuel mixture in the cylinder is configured as shown in FIG. For this reason, as shown in FIG. 7B, the G / F gradually approaches the stoichiometric air-fuel ratio from the lean as the fuel amount increases, and G / F = 14.7 at the fully open load in the SI mode. It becomes.

  FIG. 7C shows the external EGR rate. As described above, the external EGR rate becomes 0 (zero) during the execution of the internal EGR control in the CI mode, and during the control of the external EGR in the CI mode. Maximum. Thus, in the SI mode, the EGR rate gradually decreases as the load on the engine 1 increases, and becomes 0 again at the fully opened load. That is, even in the SI mode, the external EGR control is stopped at the fully open load.

  As shown in FIG. 7D, the fuel injection timing is set, for example, during the intake stroke between the exhaust top dead center and the intake bottom dead center in the CI mode. The fuel injection timing may be changed according to the load of the engine 1. On the other hand, in the SI mode, the fuel injection timing is set to a retard period from the latter half of the compression stroke to the early stage of the expansion stroke. That is, high pressure retarded injection. In the SI mode, the injection timing is gradually changed to the retard side as the engine load increases. This is because, as the engine load increases, the pressure and temperature in the cylinder 18 increase and abnormal combustion is likely to occur. Therefore, in order to avoid this efficiently, the injection timing is set to the retarded side. This is because it is necessary to set. Here, the solid line in FIG. 7D shows an example of the fuel injection timing in the case of collective injection in which the high pressure retarded injection is performed by one fuel injection. In contrast, the alternate long and short dash line in FIG. 7D shows the fuel injection for each of the first injection and the second injection when the high-pressure retarded injection is divided into two fuel injections of the first injection and the second injection. An example of timing is shown. According to this, since the second injection in the divided injection is executed on the retarded angle side compared with the case of performing the batch injection, it is advantageous for avoiding abnormal combustion. This is because, as described above, the first injection is executed relatively early to ensure the vaporization time of the fuel, and the fuel injection amount of the second injection is relatively reduced. This is because the conversion time is shortened.

  Further, as indicated by a dotted line in FIG. 7 (d), the total fuel injection amount increases in the fully open load region. Therefore, the intake stroke injection is performed for the purpose of improving the intake charge efficiency by increasing the fuel injection amount. You may make it perform.

  FIG. 7E shows a change in the fuel pressure supplied to the direct injection injector 67. In the CI mode, the fuel pressure is set constant at the minimum fuel pressure. In contrast, in the SI mode, the fuel pressure is set to be higher than the minimum fuel pressure, and the fuel pressure is set to increase as the engine load increases. This is because abnormal combustion tends to occur as the engine load increases, so that further shortening of the injection period and further retarding of the injection timing are required.

  FIG. 7 (f) shows the change in the injection pulse width (injector opening period) corresponding to the injection period in the case of batch injection. In the CI mode, the pulse width increases as the fuel injection amount increases. Similarly, in the SI mode, the pulse width increases as the fuel injection amount increases. However, as shown in FIG. 5 (e), in the SI mode, the fuel pressure is set to be significantly higher than in the CI mode. Therefore, the fuel injection amount in the SI mode is larger than the fuel injection amount in the CI mode. Nevertheless, the pulse width is set shorter than the pulse width in the CI mode. This shortens the unburned mixture reaction possible time and is advantageous for avoiding abnormal combustion.

  FIG. 7G shows a change in the ignition timing. In the SI mode, as the fuel injection timing is retarded as the engine load increases, the ignition timing also retards as the engine load increases. Is done. Thereby, at least in the fully open load region, combustion starts in the expansion stroke. This is advantageous for avoiding abnormal combustion. In the CI mode, although ignition is basically not performed, for the purpose of avoiding smoldering of the spark plug 25, for example, ignition may be performed in the vicinity of the exhaust top dead center as shown by a one-dot chain line in FIG. .

  Thus, in this engine 1, for the purpose of stabilizing compression ignition combustion in the CI mode, as shown in an enlarged view in FIGS. 8 and 9, a ceramic heater as a heating means is provided at the throat of the exhaust port 17. 81 is disposed. More specifically, the ceramic heater 81 is configured in an endless annular shape in the illustrated example, and a part of the ceramic heater 81 is embedded in the throat in each of the exhaust ports 17 formed for each cylinder 18 in the port wall surface. It is attached.

  Although the detailed structure of each ceramic heater 81 is omitted, a heating element is built in the ceramic. The heating element generates heat when energized, and the heat generation stops when energization is stopped. Switching between energization and deenergization of the ceramic heater 81 is performed by the PCM 10. The ceramic also functions as a heat accumulator and absorbs the heat of burned gas discharged from the cylinder 18 to the exhaust port 17 while being introduced into the cylinder 18 from the exhaust port 17 as will be described later. Dissipates heat to the fuel gas.

  Thus, the ceramic heater 81 is energized when the internal EGR control is executed by opening the exhaust valve 22 twice, in other words, in the low load side region in the CI mode described above. As a result, as shown in the left diagram of FIG. 10, when the burned gas is discharged from the cylinder 18 to the exhaust port 17 through the exhaust valve 22 that is opened during the exhaust stroke, the heat of the burned gas is discharged. However, heat is absorbed by the heat storage body of the ceramic heater 81. Thereafter, as shown in the right diagram of FIG. 10, when the exhaust valve 22 is opened during the intake stroke and the burned gas is introduced into the cylinder 18 from the exhaust port 17, The burnt gas is heated by the heat radiation and the heat generated by the heating element, and the burnt gas whose temperature has decreased in the exhaust port 17 is heated. In this way, the higher temperature burned gas is introduced into the cylinder 18, thereby further increasing the in-cylinder temperature. This further stabilizes the compression ignition combustion.

  When the load on the engine 1 increases and the internal EGR control is stopped, the energization to the ceramic heater 81 is also stopped. Thereby, energy loss is suppressed. On the other hand, after the energization to the ceramic heater 81 is stopped, the residual heat of the ceramic heater 81 exists, but since the opening of the exhaust valve 22 is stopped twice, the residual heat of the ceramic heater 81 has already been heated. No fuel gas is introduced into the cylinder 18. This is advantageous in avoiding the occurrence of knocking by avoiding unnecessarily increasing the temperature in the cylinder 18.

  For example, the ceramic heater 81 may be disposed at the throat of the intake port 16 instead of being disposed at the throat of the exhaust port 17 or in addition to being disposed at the throat of the exhaust port 17. The ceramic heater disposed at the throat of the intake port 16 heats the gas (that is, fresh air) introduced from the intake port 16 into the cylinder 18 during the intake stroke to raise the temperature. Therefore, the in-cylinder temperature is increased, which contributes to stabilization of compression ignition combustion. On the other hand, even after the load on the engine 1 increases and the energization is stopped, the ceramic heater disposed at the throat of the intake port 16 heats fresh air due to the residual heat. Therefore, disposing the heating means at the intake port 16 may unnecessarily increase the in-cylinder temperature. In this regard, it is preferable that the heating means be disposed only at the exhaust port 17.

  The throat of the exhaust port 17 is a portion where the gas flow rate is the highest, and disposing the ceramic heater 81 here increases the heat transfer rate, and absorbs heat from the burned gas, and dissipates heat from the burned gas and generates heat. Heating by the body can be performed efficiently.

  Further, the combination of the heating element and the heat storage material is advantageous in reducing as much energy (power consumption) required for raising the temperature of the burned gas as possible. In particular, by disposing the heat storage material (ceramic heater 81) in the throat, the temperature difference between the heat storage material and the high-temperature burned gas immediately after being discharged from the cylinder 18 is increased, and the heat absorption efficiency is increased. When the burned gas is introduced, the temperature difference between the burned gas whose temperature has decreased in the exhaust port 17 and the heat storage material is increased, so that the heat radiation efficiency is increased, which is more advantageous for increasing the temperature of the burned gas.

  Further, by disposing such a ceramic heater 81 in the exhaust port 17 and not exposing it in the cylinder 18, it is possible to avoid providing a heat source in the combustion chamber, which is advantageous in avoiding knocking.

  As described above, the engine 1 has a relatively high geometric compression ratio for the purpose of stabilizing compression ignition combustion in the CI mode. If the ignition combustion is sufficiently stabilized, the geometric compression ratio of the engine 1 can be lowered accordingly. This makes it possible to obtain merits such as a reduction in mechanical resistance loss and a reduction in cooling loss. Further, lowering the geometric compression ratio of the engine 1 makes it possible to change, for example, the arrangement of the direct injection injector 67 from the center injection shown in the figure to the side injection.

  The shape of the ceramic heater 81 is not shown as an endless ring as shown in the figure, but although not shown in the figure, it may be an endless ring or a semicircular shape with a part thereof cut off. . Further, as shown in FIG. 11, by connecting heaters in a substantially semicircular shape (a shape obtained by horizontally inverting the “C” shape) disposed corresponding to each of the two exhaust ports 17, You may form so that the character of "3" may be made as a whole. The ceramic heater 82 of this shape avoids the heater being disposed in a narrow region between the exhaust ports 17 and 17 where the ignition plug 25 is disposed, or in a narrow region between the exhaust port 17 and the intake port 16. This is advantageous for avoiding the intake side from being heated by the ceramic heater 82.

  In the above-described configuration, the ceramic heater 81 that generates heat when energized is disposed in the exhaust port, but a heat storage material (that is, a heating element is not included) may be disposed in the exhaust port instead. Conversely, a heating element (that is, not including a heat storage material) may be disposed in the exhaust port.

  In the above configuration, the ceramic heaters 81 and 82 are disposed in the two exhaust ports 17 and 17, respectively. For example, the ceramic heaters 81 and 82 are provided only in one of the two exhaust ports 17 and 17. May be arranged. Further, for example, when the exhaust valve 22 is opened twice, only one of the two exhaust valves 22 is opened, the ceramic heater 81 is connected to the exhaust port 17 corresponding to the exhaust valve 22 opened twice. Is preferably arranged.

1 Engine (Engine body)
10 PCM (controller)
16 Intake port 17 Exhaust port 18 Cylinder 22 Exhaust valve 25 Spark plug 67 Direct injection injector (fuel injection valve)
81 Ceramic heater (heating means)
82 Ceramic heater (heating means)

Claims (6)

An engine body configured to supply at least gasoline-containing fuel into the cylinder;
A controller configured to at least control the opening operation of the intake valve and the exhaust valve, and
The controller is
When the engine body is at least warmed up and its operating state is in a predetermined low load range, the burned gas is introduced into the cylinder by opening an exhaust valve during the intake stroke, and the cylinder A compression ignition mode in which the air-fuel mixture is compressed and ignited,
When the operating state of the engine body is in a region where the load is higher than that in the compression ignition mode, the introduction of the burned gas into the cylinder is substantially stopped so that the exhaust valve during the intake stroke is stopped. With the valve opening operation stopped, the engine body is operated,
Of the intake port and the exhaust port communicating with the cylinder, at least the port where the exhaust valve that opens during the intake stroke in the compression ignition mode is disposed passes through the port toward the inside of the cylinder. A heating means for heating the gas is provided ;
The heating means includes a heating element that can be switched between activated and deactivated by the controller,
The controller is a compression ignition type gasoline engine that stops the operation of the heating element when the operating state of the engine body is in the high load region . The compression ignition type gasoline engine according to claim 1,
The heating means is a compression ignition type gasoline engine provided only in the exhaust port among the intake port and the exhaust port. The compression ignition type gasoline engine according to claim 1 or 2 ,
The controller sets a spark ignition mode in which an air-fuel mixture in the cylinder is ignited by driving an ignition plug when the operating state of the engine body is at least in a fully open load region, and the ignition timing is burned in an expansion stroke. A compression ignition gasoline engine that retards to start. In the compression ignition type gasoline engine according to any one of claims 1 to 3 ,
The controller is configured so that the introduction of burned gas into the cylinder is substantially stopped when the operating state of the engine body is in a medium load region adjacent to the load region in the compression ignition mode. A compression ignition gasoline engine that compresses and ignites an air-fuel mixture in the cylinder in a state in which the valve opening operation of the exhaust valve is stopped during an intake stroke. In the compression ignition type gasoline engine according to any one of claims 1 to 4 ,
The heating means is a compression ignition type gasoline engine including a heat storage material. In the compression ignition type gasoline engine according to any one of claims 1 to 5 ,
The heating means is a compression ignition type gasoline engine provided at a throat of the port.

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