JP2020200772A - Engine control device - Google Patents

Abstract

To provide an engine control device which enables an increase in combustion noise to be surely prevented.SOLUTION: An engine comprises: a supercharger 33 which is provided in an intake passage 30 and supercharges intake air; a bypass passage 38 that is provided in the intake passage 30 and bypasses the supercharger 33; and an EGR device 50 that returns an EGR gas, which is a part of exhaust gas, to a portion of the intake passage 30 on the upstream side of the supercharger 33. An engine control device carries out combustion slowing control for slowing down the combustion of an air-fuel mixture in a cylinder 2 if an EGR rate is less than a predetermined reference EGR rate when the supercharger 33 is started to be driven, and prohibits the carrying-out of the combustion slowing control if the EGR rate is equal to or higher than the reference EGR rate.SELECTED DRAWING: Figure 7

Description

The present invention relates to an engine control device.

As disclosed in Patent Document 1, in an engine installed in a vehicle or the like, a supercharger is provided in an intake passage to increase engine torque, and a part of exhaust gas is used to suppress NOx emissions. It is recirculated to the intake passage.

Specifically, the engine of Patent Document 1 is provided with a so-called supercharger, which is a supercharger driven by the engine body 1, and a bypass passage for bypassing the supercharger in the intake passage thereof. When the supercharger is driven, air is introduced into the cylinder through the supercharger, while when the supercharger is stopped, air is introduced into the cylinder by bypassing the supercharger. Unlike a turbocharger, in which the supercharger is driven to rotate by a turbine installed in the exhaust passage, this supercharger can supercharge without increasing the back pressure of the engine, and the engine torque is effective. Can be enhanced to.

Further, the engine of Patent Document 1 is provided with an EGR device including an EGR passage connecting an exhaust passage and an intake passage and a valve for opening and closing the EGR passage. The EGR passage is connected to a portion of the intake passage on the upstream side of the supercharger, so that a part of the exhaust gas returns to the portion on the upstream side of the supercharger. When the turbocharger is driven, the pressure on the downstream side of the turbocharger becomes higher. On the other hand, in the engine of Patent Document 1, as described above, the EGR passage is connected to a portion on the upstream side of the turbocharger and a portion having a lower pressure than the portion on the downstream side. Therefore, the pressure in the exhaust passage can be surely made higher than the pressure in the intake passage, and a large amount of EGR gas can be returned to the intake passage.

Japanese Unexamined Patent Publication No. 2019-39383

When the structure of the intake system is as shown in Patent Document 1, when the supercharger is stopped, the EGR gas is mainly introduced into the cylinder through the bypass passage, and the EGR in the passage in which the supercharger is arranged is provided. The proportion of gas is low. Therefore, when the drive of the supercharger is started and the gas in the passage in which the supercharger is arranged is introduced into the cylinder, the EGR gas is temporarily not sufficiently introduced into the cylinder, and the air-fuel mixture is not sufficiently introduced. The progress of combustion of flatulence may become excessively fast. If the progress of combustion of the air-fuel mixture becomes excessively fast, the air-fuel mixture starts combustion at a desired timing, or the air-fuel mixture burns rapidly and the combustion noise increases.

The present invention has been made in view of the above circumstances, and an object of the present invention is to provide an engine control device capable of reliably preventing an increase in combustion noise.

In order to solve the above problems, the present invention provides an intake passage through which the intake air introduced into the cylinder flows, an exhaust passage through which the exhaust gas discharged from the cylinder flows, and an exhaust passage provided in the intake passage to supercharge the intake air. The supercharger, the bypass passage provided in the intake passage to bypass the supercharger, and the EGR gas which is a part of the exhaust gas in the exhaust passage are more than the supercharger in the intake passage. An engine control device equipped with an EGR device that recirculates to the upstream part, a combustion suppressing means capable of suppressing the progress of combustion of the air-fuel mixture in the cylinder, and an EGR gas in the cylinder with respect to the total amount of gas in the cylinder. A determination means for determining whether or not the EGR rate, which is a ratio of the amount, is less than a predetermined reference EGR rate, and a control means for controlling the supercharger and the combustion suppression means are provided, and the control means is stopped. When it is determined by the determination means that the EGR rate is less than the reference EGR rate at the start of supercharging for driving the supercharger in the state, the progress of combustion of the air-fuel mixture in the cylinder is suppressed. When it is determined by the determination means that the EGR rate is equal to or higher than the reference EGR rate, the execution of the combustion slowdown control is prohibited. (Claim 1).

According to this configuration, the EGR gas is returned to the portion of the intake passage on the upstream side of the supercharger, so that the amount of EGR gas introduced into the intake passage and the cylinder can be secured. Moreover, in this configuration, when the drive of the turbocharger in the stopped state is started and the EGR rate is less than the reference EGR rate and the EGR rate is low, the combustion of the air-fuel mixture in the cylinder is slowed down. Chemical control is implemented. Therefore, it is possible to prevent the combustion noise from increasing due to the air-fuel ratio of the air-fuel mixture becoming excessively lean at the start of driving the turbocharger. When the EGR rate is relatively high and the air-fuel ratio of the air-fuel mixture is not excessively lean, the combustion slowdown control that slows down the combustion is prohibited, so that the combustion becomes excessively slow. It can be avoided and the fuel efficiency can be improved.

In the above configuration, preferably, the fuel injection means that injects fuel into the cylinder and functions as the combustion suppressing means is provided, and when the combustion slowdown control is executed, the control means performs the fuel injection to the cylinder. The fuel injection means is controlled so as to be slower than when the combustion slowdown control is not executed (claim 2).

When the fuel is injected later, the mixing time and reaction time of fuel and air up to the compression top dead center are shortened, so that a large amount of air-fuel mixture with advanced reaction is generated in the cylinder near the compression top dead center. It can be prevented from being done. Therefore, according to this configuration, the reaction of the air-fuel mixture near the compression top dead center can be suppressed and the combustion can be slowed down by a simple configuration in which the timing of fuel injection is delayed.

In the above configuration, preferably, the fuel injection means performs the first fuel injection and the second fuel injection whose injection timing is later than that of the first fuel injection, and when the combustion slowdown control is executed. The control means controls the fuel injection means so that the second fuel injection is performed in the latter half of the compression stroke (claim 3).

According to this configuration, the mixing time and reaction time of the fuel and air up to the compression top dead center are surely shortened, the reaction of the air-fuel mixture up to the compression top dead center is suppressed, and the combustion is slowed down. it can.

In the above configuration, preferably, when the combustion slowdown control is executed, the control means increases the amount of air introduced into the cylinder and the amount of EGR gas introduced into the cylinder, the more the cylinder is in one combustion cycle. The fuel injection means is controlled so that the ratio of the amount of fuel supplied to the cylinder by the second fuel injection to the total amount of fuel supplied to the cylinder becomes large (claim 4).

In this configuration, the larger the amount of air introduced into the cylinder and the smaller the amount of EGR gas, the larger the proportion of the amount of fuel involved in the second fuel injection. That is, as the air-fuel ratio of the air-fuel mixture becomes larger and the combustion noise tends to increase, the amount of fuel that can shorten the mixing time and reaction time to the vicinity of compression top dead center increases. Therefore, it is possible to more reliably suppress the reaction of the air-fuel mixture up to the vicinity of the compression top dead center and slow down the combustion.

In the above configuration, preferably, when the combustion slowdown control is executed, the control means performs both the first fuel injection and the second fuel injection in the compression stroke (claim 5).

According to this configuration, the mixing time and reaction time of both the fuel related to the first fuel injection and the fuel related to the second fuel injection to the vicinity of the compression top dead center are shortened, and the progress of combustion of the air-fuel mixture is shortened. Can definitely be slowed down.

In the above configuration, preferably, there is an air filling amount changing means capable of changing the air filling amount which is the amount of air introduced into the cylinder, and the controlling means has a target EGR rate in which the EGR rate is set in advance. When the EGR rate deficiency condition that the difference between the EGR rate and the target EGR rate is greater than or equal to a predetermined determination value is satisfied, the air filling amount changing means is controlled so that the air filling amount is reduced (claimed). Item 6).

According to this configuration, the amount of air introduced into the cylinder when the shortage amount of the EGR rate with respect to the target EGR rate is large and the combustion noise tends to increase due to the air-fuel ratio of the air-fuel mixture in the cylinder being excessively large. Can be reduced to suppress an increase in combustion noise.

In the above configuration, preferably, the control means reduces the air filling amount when the EGR rate insufficient condition is satisfied and when the temperature of the intake air introduced into the cylinder is equal to or higher than a predetermined determined intake air temperature. The air filling amount changing means is controlled as described above (claim 7).

According to this configuration, when the shortage amount of the EGR rate with respect to the target EGR rate is large and the combustion noise tends to increase due to the high intake air temperature, the amount of air introduced into the cylinder is reduced to reduce the combustion noise. Can be suppressed from increasing.

In the above configuration, preferably, a purification device provided in the exhaust passage and capable of purifying NOx in the exhaust gas is provided, and the EGR device purifies the exhaust gas in the exhaust passage on the downstream side of the purification device. It is returned to the intake passage (claim 8).

The more OH radicals contained in the air-fuel mixture, the faster the progress of combustion of the air-fuel mixture. It is also known that OH radicals are generated when NOx passes through the purification device. Therefore, in an engine in which the exhaust gas in the exhaust passage on the downstream side of the purification device is returned to the intake passage and introduced into the cylinder, the OH radicals in the cylinder increase and the progress of combustion of the air-fuel mixture tends to increase. Noise tends to increase. Therefore, the above configuration includes a purification device provided in the exhaust passage and capable of purifying NOx in the exhaust gas, and the EGR device purifies the exhaust gas in the exhaust passage on the downstream side of the purification device. When applied to an engine configured to recirculate to the intake passage, an increase in combustion noise can be effectively suppressed.

The engine to which the above configuration is applied includes an ignition means for igniting the air-fuel mixture in the cylinder, and the control means sparks and burns a part of the air-fuel mixture in the cylinder by ignition from the ignition means. At the same time, the ignition means is controlled so that the remaining air-fuel mixture is burned by self-ignition (claim 9).

As described above, according to the engine control device of the present invention, an increase in combustion noise can be reliably prevented.

It is a system diagram which shows schematic the whole structure of the engine which concerns on one Embodiment of this invention. It is a block diagram which shows the control system of an engine. It is a map diagram which divided the operating area of an engine by the difference of an operating mode. It is a graph which shows the waveform of the heat generation rate at the time of SPCCI combustion. It is a figure which showed the waveform of the fuel injection pattern and the heat generation rate at the time of normal control in a 2nd operation region. It is the figure which showed the flow of the gas around the supercharger, (a) is the figure when the supercharger is stopped, (b) is the figure when the supercharger is driven. It is a flowchart which showed the flow of the combustion noise prevention control. It is a figure which showed the waveform of the fuel injection pattern and the heat generation rate at the time of carrying out the combustion noise prevention control. It is a figure which showed the relationship between the air filling amount and the EGR gas amount at the time of carrying out the combustion noise prevention control, and the fuel injection timing. It is a figure which showed the relationship between the air filling amount and EGR gas amount at the time of performing combustion noise prevention control, and the ratio of the injection amount of the post-stage injection. It is a figure which showed the relationship between the EGR rate deficiency amount, the intake air temperature and the engine water temperature, and the reduction amount of the opening degree of a throttle valve. It is a figure which showed the time change of each parameter at the time of carrying out combustion noise prevention control. It is a figure which showed the relationship between the temperature of a mixture, the equivalent ratio of a mixture, and the ease of self-ignition of the mixture. It is the figure which showed the state of the fuel spraying at the time of the pre-ign generation, (a) is the figure immediately after the execution of the pre-stage injection, (b) is the figure near the compression top dead center. It is a figure which showed the distribution of the air-fuel mixture of each temperature and each air-fuel ratio existing in a combustion chamber 6, (a) is a figure when the ratio of post-stage injection is 5%, (b) is a figure which the ratio of post-stage injection is It is a figure at the time of 30%.

(1) Overall Configuration of Engine FIG. 1 is a system diagram schematically showing the overall configuration of an engine to which the control method and control device of the engine of the present invention are applied. The engine system shown in this figure is mounted on a vehicle and includes an engine body 1 as a power source for traveling. In the present embodiment, a 4-cycle gasoline direct injection engine is used as the engine body 1. In the engine system, in addition to the engine body 1, the intake passage 30 through which the intake air introduced into the engine body 1 flows, the exhaust passage 40 through which the exhaust gas discharged from the engine body 1 flows, and the exhaust gas flowing through the exhaust passage 40 The EGR device 50 is provided so as to return a part of the EGR device to the intake passage 30.

The engine body 1 is slidable back and forth between the cylinder block 3 in which the cylinder 2 is formed, the cylinder head 4 attached to the upper surface of the cylinder block 3 so as to block the cylinder 2 from above, and the cylinder 2. It has an inserted piston 5. The engine body 1 is a multi-cylinder type having a plurality of cylinders 2 (for example, four cylinders 2 arranged in a direction orthogonal to the paper surface of FIG. 1), but here, for simplification, only one cylinder 2 is used. The explanation will proceed focusing on.

A combustion chamber 6 is defined above the piston 5, and fuel containing gasoline as a main component is supplied to the combustion chamber 6 by injection from an injector 15 described later. Then, the supplied fuel is burned while being mixed with air in the combustion chamber 6, and the piston 5 pushed down by the expansion force due to the combustion reciprocates in the vertical direction. As the fuel injected into the combustion chamber 6, a fuel containing gasoline as a main component is used. This fuel may contain auxiliary components such as bioethanol in addition to gasoline. The injector 15 is configured to be able to inject fuel in a plurality of times during one combustion cycle.

Below the piston 5, a crankshaft 7, which is an output shaft of the engine body 1, is provided. The crankshaft 7 is connected to the piston 5 via a connecting rod 8 and is rotationally driven around the central axis in response to the reciprocating motion (vertical motion) of the piston 5.

The geometric compression ratio of the cylinder 2, that is, the ratio of the volume of the combustion chamber 6 when the piston 5 is at the top dead center and the volume of the combustion chamber 6 when the piston 5 is at the bottom dead center is determined by SPCCI combustion described later. A value suitable for (partial compression ignition combustion) is set to 13 or more and 30 or less.

The cylinder block 3 is provided with a crank angle sensor SN1 that detects the rotation angle (crank angle) of the crankshaft 7 and the rotation speed (engine rotation speed) of the crankshaft 7. Further, the cylinder block 3 is provided with an engine water temperature sensor SN2 that detects the temperature of the engine cooling water, that is, the engine water temperature for circulating the water jacket formed in the cylinder block 3 to cool the engine body 1.

The cylinder head 4 is provided with an intake port 9 and an exhaust port 10 that open into the combustion chamber 6, an intake valve 11 that opens and closes the intake port 9, and an exhaust valve 12 that opens and closes the exhaust port 10. The valve type of the engine of this embodiment is a 4-valve type of 2 intake valves x 2 exhaust valves, and the intake port 9, the exhaust port 10, the intake valve 11 and the exhaust valve 12 are 2 for each cylinder 2. It is provided one by one. In the present embodiment, a swirl valve 18 that can be opened and closed is provided in one of the two intake ports 9 connected to one cylinder 2, and the swirl flow in the cylinder 2 (turns around the cylinder axis). The strength of the swirling flow) is changed.

The intake valve 11 and the exhaust valve 12 are opened and closed and driven in conjunction with the rotation of the crankshaft 7 by the valve operating mechanisms 13 and 14 including a pair of camshafts and the like arranged on the cylinder head 4.

The valve operating mechanism 13 for the intake valve 11 has a built-in intake VVT 13a capable of changing at least the opening time of the intake valve 11. Similarly, the valve operating mechanism 14 for the exhaust valve 12 includes an exhaust VVT 14a capable of changing at least the closing time of the exhaust valve 12.

In the present embodiment, by controlling the intake VVT 13a and the exhaust VVT 14a, the closing timing of the exhaust valve 12 is on the retard side of the opening timing of the intake valve 11, and both the intake valve 11 and the exhaust valve 12 are predetermined. Valve overlap that opens for a period of time has been realized. Further, by controlling the intake VVT 13a and the exhaust VVT 14a, the valve overlap period, which is the period during which both the intake valve 11 and the exhaust valve 12 are opened, is changed. When the intake valve 11 and the exhaust valve 12 are driven so as to overlap each other, the burned gas is discharged from the combustion chamber 6 to at least one of the intake passage 30 and the exhaust passage 40, and then the burned gas is discharged again. It is introduced into the combustion chamber 6. As a result, the burnt gas remains in the combustion chamber 6. The amount of burnt gas remaining in the combustion chamber 6 changes depending on the valve overlap period, and the amount of burnt gas remaining in the combustion chamber 6 is adjusted by adjusting the valve overlap period. The intake VVT 13a (exhaust VVT 14a) may be a variable mechanism of a type that changes only the closing time (opening time) while fixing the opening time (closing time) of the intake valve 11 (exhaust valve 12). It may be a phase-type variable mechanism that simultaneously changes the opening time and closing time of the intake valve 11 (exhaust valve 12).

The cylinder head 4 ignites an injector 15 that injects fuel (mainly gasoline) into the combustion chamber 6 and a mixture of fuel injected from the injector 15 into the combustion chamber 6 and air introduced into the combustion chamber 6. An ignition plug (ignition means) 16 is provided. The cylinder head 4 is further provided with a cylinder pressure sensor SN3 that detects the pressure inside the cylinder, which is the pressure of the combustion chamber 6.

The injector 15 is a multi-injection type injector having a plurality of injection holes at its tip, and can inject fuel radially from the plurality of injection holes. The injector 15 is provided so that its tip portion faces the central portion of the crown surface of the piston 5. In the present embodiment, a cavity is formed on the crown surface of the piston 5 so that the region including the central portion thereof is recessed on the opposite side (lower side) of the cylinder head 4.

The spark plug 16 is arranged at a position slightly offset from the injector 15 on the intake side.

The intake passage 30 is connected to one side surface of the cylinder head 4 so as to communicate with the intake port 9. The air (intake air, fresh air) taken in from the upstream end of the intake passage 30 is introduced into the combustion chamber 6 through the intake passage 30 and the intake port 9.

In the intake passage 30, in order from the upstream side, an air cleaner 31 for removing foreign matter contained in the intake air (air) introduced into the combustion chamber 6 (cylinder 2), a throttle valve 32, and a supercharger that compresses and sends out the intake air. A feeder 33, an intercooler 35 for cooling the intake air compressed by the supercharger 33, and a surge tank 36 are provided. The throttle valve 32 is an on-off valve that opens and closes the intake passage 30, and the air filling amount, which is the amount of air introduced into the combustion chamber 6 (cylinder 2), is changed depending on the opening degree of the throttle valve 32. To. In the present embodiment, the throttle valve 32 corresponds to the "air filling amount changing means" of the claim.

Each part of the intake passage 30 is provided with an air flow sensor SN4 for detecting the intake amount which is the flow rate of the intake air and an intake air temperature sensor SN5 for detecting the intake air temperature which is the temperature of the intake air. The air flow sensor SN4 is provided in a portion of the intake passage 30 between the air cleaner 31 and the throttle valve 32, and detects the flow rate of the intake air passing through the portion. The intake air temperature sensor SN5 is provided in the surge tank 36 and detects the temperature of the intake air in the surge tank 36.

The supercharger 33 is a mechanical supercharger (supercharger) that is mechanically linked to the engine body 1. The specific type of the turbocharger 33 is not particularly limited, but any known turbocharger such as a Rishorum type, a roots type, or a centrifugal type can be used as the supercharger 33.

An electromagnetic clutch 34 capable of electrically switching between engagement and release is interposed between the supercharger 33 and the engine body 1. When the electromagnetic clutch 34 is engaged, the driving force is transmitted from the engine body 1 to the supercharger 33, and supercharging is performed by the supercharger 33. On the other hand, when the electromagnetic clutch 34 is released, the transmission of the driving force is cut off, and supercharging by the supercharger 33 is stopped.

The intake passage 30 is provided with a bypass passage 38 for bypassing the supercharger 33. The bypass passage 38 connects the surge tank 36 and the EGR passage 51, which will be described later, to each other. A bypass valve 39 that can be opened and closed is provided in the bypass passage 38. The bypass valve 39 is a valve for adjusting the intake pressure, that is, the boost pressure, which is introduced into the surge tank 36. For example, as the opening degree of the bypass valve 39 increases, the flow rate of the intake air passing through the bypass passage 38 increases, and as a result, the boost pressure decreases.

The exhaust passage 40 is connected to the other side surface of the cylinder head 4 so as to communicate with the exhaust port 10. The burnt gas (exhaust) generated in the combustion chamber 6 is discharged to the outside through the exhaust port 10 and the exhaust passage 40.

A catalytic converter 41 is provided in the exhaust passage 40. The catalyst converter 41 includes a three-way catalyst 41a for purifying harmful components (HC, CO, NOx) contained in the exhaust gas, and a GPF (gasoline) for collecting particulate matter (PM) contained in the exhaust gas. -Particulate filter) 41b is built in from the upstream side in this order. In addition, another catalyst converter containing an appropriate catalyst such as a three-way catalyst or a NOx catalyst may be added to the downstream side of the catalyst converter 41.

The EGR device 50 has an EGR passage 51 that connects the exhaust passage 40 and the intake passage 30, and an EGR cooler 52 and an EGR valve 53 provided in the EGR passage 51. The EGR passage 51 connects a portion of the exhaust passage 40 on the downstream side of the catalytic converter 41 and a portion of the intake passage 30 between the throttle valve 32 and the supercharger 33 to each other. As a result, the EGR gas returns to the portion of the intake passage 30 on the upstream side of the supercharger 33. Specifically, in the present embodiment, the EGR passage 51 is connected to the middle part of the bypass passage 38, and the EGR gas is more than the supercharger 33 in the passage in which the supercharger 33 is arranged in the intake passage 30. It will be introduced in the upstream part. In this way, the EGR gas is configured to recirculate to the portion upstream of the supercharger 33 of the intake passage 30, so that the pressure on the intake side of the EGR passage 51 is more reliable than the pressure on the exhaust side. It gets lower. From this, a large amount of EGR gas can be recirculated. Specifically, when the supercharger 33 is driven, the pressure on the downstream side of the supercharger 33 increases. Therefore, if the EGR passage 51 is connected to the downstream portion of the turbocharger 33, the pressure on the intake side of the EGR passage 51 cannot be sufficiently lowered, and the recirculation of the EGR gas is suppressed. On the other hand, in the present embodiment, the pressure on the intake side of the EGR passage 51 can be sufficiently lowered as described above, so that the EGR gas can be reliably returned. The EGR cooler 52 cools the EGR gas, which is the exhaust gas returned from the exhaust passage 40 to the intake passage 30 through the EGR passage 51, by heat exchange. The EGR valve 53 is provided so as to be openable and closable in the EGR passage 51 on the downstream side (the side closer to the intake passage 30) of the EGR cooler 52, and adjusts the flow rate of the exhaust gas flowing through the EGR passage 51.

The EGR passage 51 is provided with a differential pressure sensor SN6 for detecting the difference between the pressure on the upstream side and the pressure on the downstream side of the EGR valve 53.

(2) Control system FIG. 2 is a block diagram showing an engine control system. The ECU 100 shown in this figure is a microprocessor for comprehensively controlling the engine, and is composed of a well-known CPU 150, a memory 160 (ROM, RAM), and the like.

Detection signals from various sensors are input to the ECU 100. For example, the ECU 100 is electrically connected to the crank angle sensor SN1, the engine water temperature sensor SN2, the in-cylinder pressure sensor SN3, the airflow sensor SN4, the intake air temperature sensor SN5, and the differential pressure sensor SN6 described above, and is detected by these sensors. Information (that is, crank angle, engine speed, engine water temperature, cylinder pressure, intake amount, intake temperature, front-rear differential pressure of EGR valve 53) is sequentially input to the ECU 100.

Further, the vehicle is provided with an accelerator sensor SN8 that detects the opening degree of the accelerator pedal operated by the driver who drives the vehicle, and the detection signal by the accelerator sensor SN8 is also input to the ECU 100.

The ECU 100 controls each part of the engine while executing various determinations, calculations, and the like based on the input signals from the respective sensors. That is, the ECU 100 is electrically connected to the intake VVT 13a, the exhaust VVT 14a, the injector 15, the spark plug 16, the swirl valve 18, the throttle valve 32, the electromagnetic clutch 34, the bypass valve 39, the EGR valve 53, and the like. A control signal is output to each of these devices based on the results of the above.

The ECU 100 corresponds to the "control means" of the claim. Further, the ECU 100 is functionally provided with an EGR rate determination unit 110 for determining whether or not the EGR rate is smaller than the lower limit EGR rate, and the EGR rate determination unit 110 is claimed. Corresponds to "judgment means".

(3) Basic Control FIG. 3 is a map diagram for explaining the difference in operation mode according to the engine speed and the engine load. As shown in this figure, the operating area of the engine is roughly divided into three operating areas, a first operating area A, a second operating area B, and a third operating area C.

The third operating region C is an region in which the engine speed is equal to or higher than the predetermined SI speed N1. The first operating region A is a region in which the engine speed is less than the SI execution speed N1 and the engine load is less than the predetermined switching load T1. The second operating region B is a residual region other than the first operating region A and the third operating region C, and is a region in which the engine speed is less than the SI implementation speed N1 and the engine load is the switching load T1 or more. is there.

In the first operating region A and the second operating region B, compression ignition combustion (hereinafter, this is referred to as SPCCI combustion), which is a mixture of SI combustion and CI combustion, is executed. In addition, "SPCCI" in SPCCI combustion is an abbreviation for "Spark Controlled Compression Ignition".

SI combustion is spark ignition combustion, in which the air-fuel mixture is ignited by a spark plug 16 and the air-fuel mixture is forcibly burned by flame propagation that expands the combustion region from the ignition point to the surroundings. is there. CI combustion is a form in which the air-fuel mixture is burned by self-ignition in an environment where the temperature and pressure are increased by the compression of the piston 5. Then, SPCCI combustion, which is a mixture of these SI combustion and CI combustion, involves SI combustion of a part of the air-fuel mixture in the combustion chamber 6 by spark ignition performed in an environment just before the air-fuel mixture self-ignites, and the SI is burned. It is a combustion mode in which the remaining air-fuel mixture in the combustion chamber 6 is CI-combusted by self-ignition after combustion (due to further increase in temperature and pressure accompanying SI combustion).

FIG. 4 is a graph showing changes in the heat generation rate (J / deg) and changes in the amount of heat generation with respect to the crank angle when SPCCI combustion occurs. In SPCCI combustion, the heat generation during SI combustion is gentler than the heat generation during CI combustion. For example, as shown in FIG. 4, the waveform of the heat generation rate when SPCCI combustion is performed has a relatively small rising slope. Further, the pressure fluctuation in the combustion chamber 6 (that is, dP / dθ: P is the in-cylinder pressure θ is the crank angle) is also gentler during SI combustion than during CI combustion. In other words, the waveform of the heat generation rate during SPCCI combustion is relative to the first heat generation rate part (the part indicated by M1) formed by SI combustion and having a relatively small rising slope. The second heat generating portion (the portion indicated by M2) having a large rising slope is formed so as to be continuous in this order.

When the temperature and pressure in the combustion chamber 6 increase due to SI combustion, the unburned air-fuel mixture self-ignites and CI combustion is started. As illustrated in FIG. 4, the slope of the waveform of the heat generation rate changes from small to large at the timing of this self-ignition (that is, the timing when CI combustion starts). That is, the waveform of the heat generation rate in SPCCI combustion has an inflection point (X in FIG. 4) that appears at the timing when CI combustion starts.

After the start of CI combustion, SI combustion and CI combustion are performed in parallel. Since CI combustion generates more heat than SI combustion, the heat generation rate is relatively large. However, since CI combustion is performed after the compression top dead center, the slope of the heat generation rate waveform does not become excessive. That is, when the compression top dead center is passed, the motoring pressure decreases due to the lowering of the piston 5, and as a result of suppressing the increase in the heat generation rate, it is possible to avoid excessive dP / dθ during CI combustion. To. In this way, in SPCCI combustion, CI combustion is performed after SI combustion, so that dP / dθ, which is an index of combustion noise, is unlikely to become excessive, and simple CI combustion (when all fuels are CI-combusted). ), Combustion noise can be suppressed.

With the end of CI combustion, SPCCI combustion also ends. Since CI combustion has a faster combustion rate than SI combustion, the combustion end time can be earlier than that of simple SI combustion (when all fuels are SI-combusted). In other words, in SPCCI combustion, the combustion end time can be brought closer to the compression top dead center within the expansion stroke. As a result, in SPCCI combustion, fuel efficiency can be improved as compared with simple SI combustion.

(A) First operating area A
In the first operating region A, SPCCI combustion is performed while the air-fuel ratio (A / F) in the combustion chamber 6 is made higher (lean) than the stoichiometric air-fuel ratio in order to improve fuel efficiency. In the first operating region A, the air-fuel ratio in the combustion chamber 6 is increased to such an extent that the amount of rawNOx, which is NOx generated in the combustion chamber 6, is sufficiently small. For example, in the first operating region A, the air-fuel ratio in the combustion chamber 6 is set to about 30.

In the first operating region A, each part of the engine is driven as follows so that the above-mentioned combustion is realized.

In the first operating region A, the injector 15 injects into the combustion chamber 6 an amount of fuel such that the air-fuel ratio (A / F) in the combustion chamber 6 is higher than the theoretical air-fuel ratio as described above. In the present embodiment, the injector 15 is driven so that almost all of the fuel to be supplied to the combustion chamber 6 during one combustion cycle is injected into the combustion chamber 6 during the intake stroke. For example, in the first operating region A, most of the fuel is injected during the intake stroke, and the remaining fuel is injected in two steps during the compression stroke.

In the first operating region A, the spark plug 16 ignites the air-fuel mixture near the compression top dead center. SPCCI combustion is started with this ignition as a trigger, and a part of the air-fuel mixture in the combustion chamber 6 is burned by flame propagation (SI combustion), and then the remaining air-fuel mixture is burned by self-ignition (CI combustion). In addition, in order to activate the air-fuel mixture, additional ignition may be performed before the ignition performed near the compression top dead center.

In the first operating region A, the opening degree of the throttle valve 32 is set to be fully open or close to fully open.

In the first operating region A, the EGR valve 53 is fully closed, and the amount of EGR gas introduced into the combustion chamber 6 via the EGR passage 51 is set to zero.

If the high-temperature burnt gas remains in the combustion chamber 6, the temperature of the air-fuel mixture is raised, so that the air-fuel mixture can be appropriately CI-combusted. From this, in the first operating region A, the intake VVT 13a and the exhaust VVT 14a drive the intake valve 11 and the exhaust valve 12 so that they overlap each other. In the present embodiment, the intake valve 11 and the exhaust valve 12 are driven so as to open the valve for a predetermined period across the exhaust top dead center.

In the first operating region A, the swirl valve 18 is fully closed or closed to a low opening close to fully closed.

In the first operating region A, the driving of the supercharger 33 is stopped. That is, the electromagnetic clutch 34 is released to release the connection between the supercharger 33 and the engine body 1, and the bypass valve 39 is fully opened to stop the supercharging by the supercharger 33.

(B) Second operating region In a region where the engine load is high, it becomes difficult to make the air-fuel ratio of the air-fuel mixture lean because the amount of fuel supplied into the combustion chamber 6 is large. As a result, in the second operating region B, where the engine load is higher than that of the first operating region A, the air-fuel mixture is SPCCI-combusted while keeping the air-fuel ratio in the combustion chamber 6 below the stoichiometric air-fuel ratio. In the present embodiment, in the second operating region B, the air-fuel ratio of the air-fuel mixture is substantially the stoichiometric air-fuel ratio.

In the second operating region B, the opening degree of the throttle valve 32 is set so that the amount of air corresponding to the engine load is introduced into the combustion chamber 6.

In the second operating region B, the injector 15 injects into the combustion chamber 6 an amount of fuel such that the air-fuel ratio becomes the stoichiometric air-fuel ratio as described above. In the present embodiment, the injector 15 is basically the fuel to be driven into the combustion chamber 6 during one combustion cycle, as in the first operation region A, except when the pre-ign avoidance pattern is set as described later. Most of the fuel is injected during the intake stroke. Further, the injector 15 basically injects the remaining fuel from the latter half of the intake stroke to the first half of the compression stroke, except when the injection pattern is changed to the pre-ign avoidance pattern described later. For example, in the second operating region B, as shown in FIG. 5, the first fuel injection (Q1) is started near the center of the compression stroke (300 ° CA before compression top dead center, etc.), and the intake bottom dead center (intake bottom dead center). The second fuel injection (Q2) is started at BDC, 180 ° CA before compression top dead center. The injection amount of the second fuel injection (mass of the fuel injected from the injector 15 at the first time) is smaller than the injection amount of the first fuel injection (mass of the fuel injected from the injector 15 at the second time). For example, the ratio of the injection amount of the second fuel injection to the total amount of fuel injected into the combustion chamber 6 in one combustion cycle is about 10%. In the following, when the injector 15 injects fuel in two steps as described above, the first fuel injection is referred to as a pre-stage injection, and the second fuel injection is referred to as a post-stage injection.

Also in the second operating region B, the spark plug 16 ignites the air-fuel mixture near the compression top dead center. In the second operating region B as well, SPCCI combustion is started by this ignition, a part of the air-fuel mixture in the combustion chamber 6 is burned by flame propagation (SI combustion), and then the remaining air-fuel mixture is self-ignited. Combustion (CI combustion).

In the second operating region B, the EGR valve 53 is opened and the EGR gas is introduced into the combustion chamber 6 through the EGR passage 51 in order to reduce the NOx generated in the combustion chamber 6. However, when the engine load is high, a large amount of air must be introduced into the combustion chamber 6, so it is necessary to reduce the amount of EGR gas introduced into the combustion chamber 6. From this, in the second operating region B, the opening degree of the EGR valve 53 is controlled so that the amount of EGR gas introduced into the combustion chamber 6 decreases as the load side increases, and the EGR valve in the region where the engine load becomes maximum. 53 is fully closed.

Also in the second operating region B, the intake VVT 13a and the exhaust VVT 14a drive the intake valve 11 and the exhaust valve 12 so that the intake valve 11 and the exhaust valve 12 overlap each other.

In the second operating region B, the swirl valve 18 is opened to an appropriate intermediate opening degree excluding fully closed / fully opened, and the opening degree is increased as the engine load is higher.

The supercharger 33 is stopped on the side of the second operating region B where both the engine speed and the engine load are low. On the other hand, in the other area of the second operating area B, the supercharger 33 is operated. That is, the electromagnetic clutch 34 is engaged, and the supercharger 33 and the engine body 1 are connected. At this time, the opening degree of the bypass valve 39 is controlled so that the pressure (supercharging pressure) in the surge tank 36 matches the target pressure predetermined for each operating condition (engine speed / engine load).

(C) Third operating region In the third operating region C, relatively orthodox SI combustion is executed. In order to realize this SI combustion, in the third operating region C, the injector 15 injects fuel for a predetermined period at least overlapping the intake stroke. The spark plug 16 ignites the air-fuel mixture near the compression top dead center. In the third operating region C, SI combustion is started triggered by this ignition, and all of the air-fuel mixture in the combustion chamber 6 is burned by flame propagation.

In the third operating area C, the supercharger 33 is operated. The throttle valve 32 is fully opened. The opening degree of the EGR valve 53 is controlled so that the air-fuel ratio in the combustion chamber 6 is equal to or less than the stoichiometric air-fuel ratio. For example, in the third operating region C, the opening degree of the EGR valve 53 is controlled so that the air-fuel ratio in the combustion chamber 6 becomes the stoichiometric air-fuel ratio or slightly smaller than this. In the third operating region C, the swirl valve 18 is fully opened.

(4) Combustion Noise Increase Prevention Control As described above, the supercharger 33 is stopped in the first operating region A, and the supercharger 33 is driven in most of the second operating region B. Therefore, the amount of air introduced into the combustion chamber 6 is excessive at the start of supercharging when the driving of the supercharger 33 in the stopped state is started with the transition from the first operating area A to the second operating area B. As a result, the air-fuel ratio of the air-fuel mixture tends to become excessively lean (large). If the air-fuel ratio becomes excessively lean, combustion noise may increase.

This will be specifically described with reference to FIGS. 6A and 6B. 6 (a) and 6 (b) are diagrams schematically showing the flow of EGR gas around the supercharger 33 in the intake passage 30, and FIG. 6 (a) is a diagram when the supercharger 33 is stopped. 6 (b) is a diagram at the start of driving the supercharger 33. As shown in FIG. 6A, while the turbocharger 33 is stopped, the EGR gas passes through the EGR passage 51 and then mainly flows into the combustion chamber 6 through the bypass passage 38. As a result, the amount of EGR gas introduced into the intake passage 30 in which the supercharger 33 is arranged is reduced, and the gas FA in the intake passage 30 in which the supercharger 33 is arranged is the amount of EGR gas. The gas has a low proportion and a high proportion of fresh air. When the driving of the supercharger 33 is started in this state, the main flow of gas is switched to the passage side where the supercharger 33 is arranged. In particular, when the bypass valve 39 is fully closed with the start of driving the supercharger 33, the flow of gas to the combustion chamber 6 through the bypass passage 38 is stopped. Therefore, as shown in FIG. 6B, the gas FA that exists in the intake passage 30 in which the supercharger 33 is arranged and has a low proportion of EGR gas is first introduced into the combustion chamber 6. Therefore, when the driving of the turbocharger 33 in the stopped state is started, the amount of EGR gas introduced into the combustion chamber 6 becomes small, and the air-fuel ratio of the air-fuel mixture may become excessively lean.

Further, even when the EGR gas introduced into the combustion chamber 6 is insufficient regardless of the driving state of the supercharger 33, the air-fuel ratio of the air-fuel mixture in the combustion chamber 6 may become excessively lean.

When the air-fuel ratio of the air-fuel mixture in the combustion chamber 6 becomes lean, the reaction of the air-fuel mixture is promoted and the progress of combustion becomes faster. When the progress of combustion becomes high, the air-fuel mixture starts combustion at a timing earlier than the desired timing, or the combustion speed of the air-fuel mixture becomes excessively high, and the in-cylinder pressure rises sharply. Therefore, if the air-fuel ratio of the air-fuel mixture becomes excessively lean, the combustion noise may increase.

Therefore, in the present embodiment, when the combustion noise may increase due to the start of driving of the turbocharger 33 or the shortage of EGR gas, the combustion of the air-fuel mixture is slowed down in order to prevent the increase in the combustion noise. Enforce control.

The control for preventing the increase in combustion noise performed by the ECU 100 will be described with reference to the flowchart of FIG.

First, in step S1, the ECU 100 determines whether or not the driving of the supercharger 33 in the stopped state has started (whether or not the supercharger 33 has changed from OFF to ON). That is, it is determined whether or not the electromagnetic clutch 34 is engaged and the supercharger 33 and the engine body 1 are connected.

If the determination in step S1 is NO and it is not the start of driving the supercharger, the process proceeds to step S4. On the other hand, if the determination in step S1 is YES and the driving of the turbocharger in the stopped state is started, the process proceeds to step S2.

In step S2, the ECU 100 determines whether or not the EGR rate of the air-fuel mixture in the combustion chamber 6 is less than a preset lower limit EGR rate. In the present embodiment, the determination in step S2 is performed by the EGR rate determination unit 110. The EGR ratio of the air-fuel mixture is the ratio of the amount (mass) of EGR gas in the combustion chamber 6 to the total amount of gas in the combustion chamber 6, that is, the total amount (mass) of gas in the combustion chamber 6. The ECU 100 estimates the EGR rate of the air-fuel mixture based on the intake air amount detected by the airflow sensor SN4 and the front-rear differential pressure of the EGR valve 53 detected by the differential pressure sensor SN6. In step S2, it is determined whether or not the estimated EGR rate (hereinafter, appropriately referred to as an estimated EGR rate) is less than the lower limit EGR rate. In the present embodiment, the lower limit EGR rate is preset in the engine speed, engine load, intake air temperature, and engine water temperature and stored in the ECU 100. The ECU 100 extracts values corresponding to the current engine speed, engine load, intake air temperature, and engine water temperature and uses them for the determination in step S2. The lower limit EGR rate corresponds to the "reference EGR rate" of the claim.

If the determination in step S2 is NO and the EGR rate is equal to or greater than the lower limit EGR rate, the process proceeds to step S4. On the other hand, if the determination in step S2 is YES and the EGR rate is less than the lower limit EGR rate, the process proceeds to step S3.

In step S3, the ECU 100 retards the fuel injection, changes the fuel injection pattern to the pre-ign avoidance pattern, and injects fuel into the injector 15 in this pre-ign avoidance pattern.

Specifically, as shown by the broken line in FIG. 8, and as described above, the fuel injection pattern during normal control in the second operating region B (when step S3 is not performed) is the pre-stage injection (Q1). Is started during the intake stroke, and the post-stage injection (Q2) is started from the latter half of the intake stroke to the first half of the compression stroke. On the other hand, the pre-ign avoidance pattern is a pattern in which the first-stage injection (Q11) is started in the first half of the compression stroke and the second-stage injection (Q12) is started in the second half of the compression stroke, as shown by the solid line in FIG. When the pre-igu avoidance pattern is executed, the execution time (injection timing) of both the pre-stage injection and the post-stage injection is retarded with respect to the normal control time (when the pre-igu avoidance pattern is not executed). Further, in the pre-ign avoidance pattern, the pre-stage injection (Q11) is set to be on the retard side of the valve closing time IVC of the intake valve 11. For example, in the pre-ign avoidance pattern, the pre-stage injection (Q11) is started at about 100 ° CA before the compression top dead center, and the post-stage injection (Q12) is started at about 5 to 20 ° CA before the compression top dead center. The pre-stage injection (Q11) in this pre-ig avoidance pattern corresponds to the "first fuel injection" of the claim, and the post-stage injection (Q12) in the pre-igu avoidance pattern corresponds to the "second fuel injection" of the claim. Equivalent to.

In the present embodiment, the start timings of the pre-stage injection (Q11) and the post-stage injection (Q12) in the pre-ign avoidance pattern are the air filling amount, which is the amount of air introduced into the combustion chamber 6, and the EGR gas amount in the combustion chamber 6. , Set based on intake air temperature and engine water temperature. Specifically, as shown in FIG. 9, the larger the air filling amount and the smaller the EGR gas amount, the later the start time of each injection (Q11, Q12). Further, the higher the intake air temperature, the later the start time of each injection (Q11, Q12), and the higher the engine water temperature, the later the start time of each injection (Q11, Q12). The ECU 100 estimates the air filling amount from the detected value of the air flow sensor SN4 and the like, and estimates the EGR gas amount in the combustion chamber 6 from this and the estimated EGR rate and the like.

Further, in the pre-ign avoidance pattern, the ratio of the injection amount of the subsequent injection to the total amount of fuel injected into the combustion chamber 6 during one combustion cycle is larger than that in the normal control. For example, the ratio of the injection amount of the post-stage injection during normal control is about 10%, whereas it is about 20 to 40% at the time of pre-ign determination.

In the present embodiment, the ratio of the injection amount of the subsequent injection in the pre-ign avoidance pattern is set based on the air filling amount, the EGR gas amount in the combustion chamber 6, the intake air temperature, and the engine water temperature. Specifically, as shown in FIG. 10, the larger the air filling amount and the smaller the EGR gas amount, the larger the ratio of the injection amount of the post-stage injection. Further, the higher the intake air temperature, the larger the ratio of the injection amount of the latter stage injection, and the higher the engine water temperature, the larger the ratio of the injection amount of the second stage injection.

The process of step S3 is carried out from the time when the determination in step S1 is YES and the determination in step S2 is YES and the process in step S3 is started until the determination in step S2 is NO. That is, the control of step S3 for retarding the execution timing (injection timing) of each fuel injection is started when the driving of the turbocharger 33 is started in a state where the EGR rate is less than the lower limit EGR rate, and the EGR is started. It is continued until the rate becomes equal to or higher than the lower limit EGR rate.

In the present embodiment, the control of step S3, that is, the control of retarding the fuel injection execution time and changing the fuel injection pattern to the priming avoidance pattern corresponds to the "combustion slowdown control" of the claim. .. Then, the injector 15 that realizes this pre-ign avoidance pattern and slows down the combustion of the air-fuel mixture corresponds to the "fuel injection means" and the "combustion suppressing means" of the claims.

After step S3, the process proceeds to step S4. In step S4, the ECU 100 determines whether or not the insufficient amount of the EGR rate with respect to the target EGR rate, which is the target value of the EGR rate, is equal to or greater than the predetermined insufficient determination amount. That is, it is determined whether or not the EGR rate deficiency condition that the deficiency amount of the EGR rate is equal to or more than the predetermined determination deficiency amount is satisfied. The estimated EGR rate is used for this determination. The target EGR rate is preset and stored in the ECU 100. For example, the target EGR rate is preset in accordance with the engine speed, engine load, etc. and stored in the ECU 100 as a map, and the ECU 100 extracts values corresponding to the current engine speed and engine load from this map. To do. Further, the determination shortage amount is also preset and stored in the ECU 100. In the following, the shortage amount of the EGR rate with respect to the target EGR rate is simply referred to as the shortage amount of the EGR rate. The above-mentioned determination shortage amount corresponds to the "determination value" of the claim.

If the determination in step S4 is NO and the insufficient amount of the EGR rate is less than the insufficient determination amount, the process ends as it is (returns to step S1). On the other hand, if the determination in step S4 is YES and the insufficient amount of the EGR rate is equal to or greater than the insufficient determination amount, the process proceeds to step S5.

In step S5, the ECU 100 determines whether or not the condition that the intake air temperature is equal to or higher than the predetermined determination intake air temperature or the engine water temperature is equal to or higher than the predetermined determination water temperature is satisfied. The determination intake air temperature and the determination water temperature are preset and stored in the ECU 100.

If the determination in step S5 is NO, the intake air temperature is lower than the determination intake air temperature, and the engine water temperature is less than the determination water temperature, the process ends as it is (returns to step S1).

On the other hand, if the determination in step S5 is YES and the intake air temperature is equal to or higher than the determination intake air temperature, or the engine water temperature is equal to or higher than the determination water temperature, the process proceeds to step S6.

In step S6, the ECU 100 reduces the air filling amount. Specifically, the ECU 100 reduces the opening degree of the throttle valve 32 (closed side). At this time, as shown in FIG. 11, the ECU 100 increases the opening degree reduction amount of the throttle valve 32 when the insufficient amount of the EGR rate is large. Further, in the ECU 100, the higher the intake air temperature and the engine water temperature, the larger the amount of reduction in the opening degree of the throttle valve 32. After step S6, the process ends (returns to step S1).

In this way, in the present embodiment, at the start of driving the turbocharger 33, when the EGR rate is low, the fuel injection execution time is retarded, and when the EGR rate is high, the fuel injection retard angle control is performed. It will be stopped. Then, in each operating condition including the start of driving the supercharger 33, when the insufficient amount of the EGR rate is large, the air filling amount is reduced. However, in the present embodiment, even when the insufficient amount of the EGR rate is large, when the intake air temperature and the engine water temperature are low, the control for reducing the air filling amount is not implemented, and the opening degree of the throttle valve 32 is normally controlled. It is set to the opening time (when step S6 is not performed).

FIG. 12 shows the time change of each parameter before and after the start of driving of the turbocharger 33 when the above control is performed. In FIG. 12, in order from the top, the accelerator opening, the rotation speed of the supercharger 33, the air filling amount, the amount of EGR gas in the combustion chamber 6, the fuel injection execution time (fuel injection time), and the throttle valve 32. Each graph of the opening degree is shown.

When the accelerator pedal is depressed at time t1 and the operation point shifts to the operation point in the second operating area B, the supercharger 33 starts to be driven at time t2 and the rotation speed of the supercharger 33 increases. To go. By starting the driving of the turbocharger 33 in this way, at time t2, the air filling amount starts to increase and the EGR gas amount starts to decrease. Then, at the time t3 when the EGR gas amount decreases and the EGR rate becomes less than the lower limit EGR rate, the fuel injection execution time is delayed. Until time t5, the amount of retardation of the fuel injection timing increases as the amount of EGR gas decreases. After the time t5, the amount of EGR gas starts to increase, and the amount of retardation of the fuel injection timing is reduced accordingly. Then, when the EGR rate exceeds the lower limit EGR rate at time t6, the retard angle of the fuel injection timing ends. Further, when the amount of EGR gas continues to decrease and the insufficient amount of EGR rate becomes equal to or greater than the insufficient amount of determination at time t4, the opening degree of the throttle valve 32 is reduced. Then, after time t5, the amount of reduction in the opening degree of the throttle valve 32 is reduced as the amount of EGR gas increases. In the example of FIG. 12, after a while after the time t6, the opening degree of the throttle valve 32 returns to the opening degree at the time of normal control as the insufficient amount of the EGR rate becomes less than the determination insufficient amount.

(5) Action, etc. As described above, in the present embodiment, when the driving of the turbocharger 33 is started and the EGR rate is less than the lower limit EGR rate, the fuel injection pattern is changed to the pre-ign avoidance pattern. Therefore, the injection timing of fuel injection is delayed. Therefore, even if the EGR rate becomes less than the lower limit EGR rate at the start of driving the turbocharger 33 and the air-fuel ratio of the air-fuel mixture becomes excessively lean, an increase in combustion noise can be suppressed. Specifically, when the fuel is injected later, the mixing time and reaction time of the fuel and air up to the compression top dead center become shorter. Therefore, it is possible to prevent the occurrence of a situation in which a large amount of the air-fuel mixture in which the reaction has proceeded exists in the combustion chamber 6 near the compression top dead center. Therefore, the progress of combustion of the air-fuel mixture near the compression top dead center can be slowed down. That is, by excessively promoting the reaction of the air-fuel mixture near the compression top dead center, it is possible to prevent the air-fuel mixture from starting combustion at an excessively early timing or the combustion speed of the air-fuel mixture becoming excessively high. Therefore, it is possible to avoid a sudden rise in pressure in the combustion chamber 6 and suppress an increase in combustion noise.

On the other hand, in the present embodiment, even when the driving of the turbocharger 33 is started, if the EGR rate is equal to or higher than the lower limit EGR rate and the air-fuel ratio of the air-fuel mixture is not excessively lean, fuel injection is performed. The control that retards the implementation time of is stopped. Therefore, when the air-fuel ratio is not excessively lean and the increase in combustion noise can be avoided, the progress of combustion can be avoided from being slowed down, and the air-fuel mixture can be appropriately burned.

Further, in the present embodiment, in the preig avoidance pattern, the execution time of the post-stage injection (Q12) is set to the latter half of the compression stroke. Therefore, the mixing time and reaction time of the fuel and air up to the compression top dead center can be surely shortened, and the progress of combustion of the air-fuel mixture can be surely slowed down.

Further, in the pre-ign avoidance pattern, the larger the air filling amount and the smaller the amount of EGR gas introduced into the cylinder, the larger the ratio of the injection amount of the post-stage injection. Therefore, when the amount of air filling is large and the amount of EGR gas introduced into the cylinder is small and the progress of combustion tends to be fast, the mixing time and reaction to the vicinity of the compression top dead center of more fuel You can shorten the time. Therefore, the progress of combustion of the air-fuel mixture near the compression top dead center can be more reliably slowed down.

Further, in the pre-ign avoidance pattern, both the pre-stage injection (Q11) and the post-stage injection (Q12) are performed in the compression stroke. Therefore, for all fuels, the mixing and reaction time with air can be shortened, and the progress of combustion of the air-fuel mixture can be more reliably slowed down. In particular, in the present embodiment, the pre-stage injection (Q11) is performed after the intake valve 11 is closed. Therefore, the progress of combustion of the air-fuel mixture can be more reliably slowed down. The effect of setting the execution time of the first-stage injection (Q11) and the second-stage injection (Q12) as described above will be specifically described with reference to FIG. 13 and the like.

FIG. 13 shows the temperature inside the cylinder, which is the temperature inside the combustion chamber 6, that is, the temperature of the air-fuel mixture in the combustion chamber 6, the equivalent ratio of the air-fuel mixture (the value obtained by dividing the stoichiometric air-fuel ratio by the air-fuel ratio), and the self-ignition of the air-fuel mixture. This is the result of investigating the relationship with ease of use. In the graph of FIG. 13, the horizontal axis represents the temperature inside the cylinder, and the vertical axis represents the ease of self-ignition. Specifically, the vertical axis is the amount of OH radicals required to self-ignite the air-fuel mixture, and when this amount is large and a large amount of OH radicals are required to self-ignite the air-fuel mixture, the air-fuel mixture Can be said to be difficult to self-ignite. The three lines in the graph of FIG. 13 are lines in which the equivalent ratios (air-fuel ratios) of the air-fuel mixture are different from each other, and the lower lines have a larger equivalent ratio and a smaller (richer) air-fuel ratio.

As shown in FIG. 13, the air-fuel mixture is more likely to self-ignite as the temperature inside the cylinder, that is, the temperature of the air-fuel mixture is higher. Here, the central portion of the combustion chamber 6 has a higher temperature than the outer peripheral portion. From this, the self-ignition of the air-fuel mixture first starts in the central portion of the combustion chamber 6. On the other hand, in the present embodiment, the combustion of the air-fuel mixture in the central portion of the combustion chamber 6 is excessively promoted by performing the first-stage injection (Q11) and the second-stage injection (Q12) at the above-mentioned timing. Is prevented from being done.

Specifically, in the first half of the compression stroke and after the intake valve 11 is closed, the intake flow in the combustion chamber 6 is weak as the piston 5 is rising and the intake valve 11 is closed. Therefore, the pre-stage injection (Q11) is performed in the first half of the compression stroke and after the intake valve 11 is closed, so that the diffusion of the fuel F1 injected by the pre-stage injection (Q11) is as shown in FIG. 14 (a). It is suppressed and most of this fuel F1 stays near the outer periphery of the combustion chamber 6. Then, as shown in FIG. 14B, this fuel distribution is generally maintained up to the vicinity of compression top dead center. Therefore, the mixing and reaction of fuel and air related to the pre-stage injection (Q11) mainly occurs near the outer periphery of the combustion chamber 6, and the amount of temperature rise near the center of the combustion chamber 6 accompanying this reaction is suppressed to a small value. Further, the amount of the air-fuel mixture in the state immediately before self-ignition distributed in the center of the combustion chamber 6 near the compression top dead center can be suppressed to a small amount. Therefore, the timing of self-ignition of the air-fuel mixture starting near the center of the combustion chamber 6 is delayed, and the progress of combustion is delayed. During normal control (other than when combustion noise prevention control is performed), the fuel of the front stage injection is diffused to the entire combustion chamber 6 because the front stage injection (Q1) is performed during the intake stroke.

Here, in the pre-ign avoidance pattern, the post-stage injection is performed in the latter half of the compression stroke. When fuel is injected at a timing close to the compression top dead center in the latter half of the compression stroke, the fuel is supplied to the vicinity of the center of the combustion chamber 6. Further, in the pre-ign avoidance pattern, the ratio of the injection amount of the post-stage injection is increased. From this, in the pre-ignition avoidance pattern, the air-fuel ratio of the air-fuel mixture near the center of the combustion chamber 6 becomes small (rich) near the compression top dead center, and from the point of view of the air-fuel ratio, the air-fuel mixture is near the center of the combustion chamber 6. It can be said that it becomes easier to self-ignite. However, as described above, in the pre-ign avoidance pattern, the amount of temperature rise near the center of the combustion chamber 6 due to the pre-stage injection can be suppressed to a small value. Further, as described above, since the time from the post-stage injection to the compression top dead center is short, the reaction between the fuel and air related to the post-stage injection is not sufficiently advanced in the vicinity of the compression top dead center, which also causes the combustion chamber 6 The temperature rise near the center of the is suppressed. Therefore, the self-ignition of the air-fuel mixture is not excessively promoted near the center of the combustion chamber 6, and the progress of combustion of the air-fuel mixture is slowed down.

This is clear from the comparison between FIGS. 15 (a) and 15 (b). 15 (a) and 15 (b) are diagrams showing the distribution of the air-fuel mixture of each temperature and each air-fuel ratio existing in the combustion chamber 6 at the compression top dead center. In FIGS. 15A and 15B, the darker the color, the larger the distribution ratio of the air-fuel mixture. Further, FIG. 15A is a diagram when the ratio of the post-stage injection is 5% (when the ratio of the pre-stage injection is 95%), and FIG. 15B is a diagram when the ratio of the post-stage injection is 30%. It is a figure at the time of (when the ratio of the pre-stage injection is 70%). The broken lines shown in FIGS. 15 (a) and 15 (b) are lines connecting the temperature and the air-fuel ratio at which the ease of self-ignition is the same. As shown in FIG. 15A, when the ratio of the pre-stage injection is large and the ratio of the post-stage injection is small, the air-fuel ratio is relatively large (lean), but the ratio of the air-fuel mixture having a high temperature is large. On the other hand, as shown in FIG. 15B, when the ratio of the first stage injection is small and the ratio of the second stage injection is large, the air-fuel ratio is relatively small (rich) but the ratio of the air-fuel mixture is low. Becomes larger. Then, more air-fuel mixture is distributed in the region where self-ignition is difficult to occur in the case of FIG. 15 (b) than in the case of FIG. 15 (a). Therefore, in the pre-ignition avoidance pattern, although the ratio of the post-stage injection is increased and an air-fuel mixture having a high air-fuel ratio is formed near the center of the combustion chamber 6, the air-fuel mixture is in a state of being difficult to self-ignite and the combustion of the air-fuel mixture is burned. Progress is slowed down.

Further, in the present embodiment, when the EGR rate insufficient condition that the EGR rate insufficient amount is equal to or greater than the determination insufficient amount is satisfied, the opening degree of the throttle valve 32 is reduced and the air filling amount is reduced. Therefore, when the shortage amount of the EGR rate with respect to the target EGR rate is large and the combustion noise tends to increase as the air-fuel ratio of the air-fuel mixture in the combustion chamber 6 is larger than the appropriate value, the air filling amount is effectively increased. It can be reduced and the increase in combustion noise can be reliably prevented.

In particular, in the present embodiment, the air filling amount is reduced when the EGR rate insufficient condition is satisfied and when the intake air temperature is equal to or higher than the determined intake air temperature. Therefore, when the shortage amount of the EGR rate with respect to the target EGR rate is large and the combustion noise tends to increase due to the high intake air temperature, the amount of air introduced into the combustion chamber 6 is reduced and the combustion noise increases. Is effectively suppressed. Then, it is possible to prevent the opportunity for the air filling amount to be reduced to become excessively large and secure the engine torque.

Here, the more OH radicals contained in the air-fuel mixture, the more the combustion of the air-fuel mixture is promoted. It is also known that OH radicals are generated when NOx passes through the purification device. As a result, in the engine of the present embodiment configured so that the exhaust gas in the exhaust passage 40 on the downstream side of the three-way catalyst 41a is returned to the intake passage 30 and introduced into the cylinder, the OH radical in the cylinder 2 is generated. As the number increases, the progress of combustion of the air-fuel mixture becomes faster, and combustion noise tends to increase. On the other hand, in the present embodiment, as described above, the throttle valve 32 and the injector 15 are controlled so as to suppress the increase in combustion noise. Therefore, the increase in combustion noise can be effectively suppressed.

(6) Modification Example In the above embodiment, the case where the engine is subjected to SPCCI combustion has been described, but the combustion embodiment carried out by the engine is not limited to this.

Further, the operating area in which the supercharger 33 is driven is not limited to the above.

Further, in the above embodiment, the case where the EGR passage 51 is connected to the middle portion of the bypass passage 38 has been described, but in the EGR device 50, the EGR gas is on the upstream side of the supercharger 33 of the intake passage 30. The connection point of the EGR passage 51 is not limited to the above as long as it is configured to be recirculated to the portion. For example, the EGR passage 51 is connected to a position upstream of the upstream end of the bypass passage 38 in the intake passage 30 or a position between the upstream end of the bypass passage 38 and the supercharger 33 in the intake passage 30. You may. Even in this case, when the supercharger 33 is stopped, the EGR gas is mainly introduced into the combustion chamber 6 through the bypass passage 38, so that the proportion of the EGR gas in the portion upstream of the supercharger 33 becomes small. At the start of driving the supercharger 33, the air-fuel mixture in the combustion chamber 6 may become excessively lean. From this, even in the above case, if the control according to the above embodiment is performed, the increase in combustion noise can be surely suppressed.

1 Engine body 2 Cylinders 6 Combustion chamber 15 Injectors (fuel injection means, combustion suppression means)
30 Intake passage 32 Throttle valve (means for changing air filling amount)
33 Supercharger 38 Bypass passage 40 Exhaust passage 50 EGR device 100 ECU (control means)

Claims (9)

An intake passage through which the intake air introduced into the cylinder flows, an exhaust passage through which the exhaust gas discharged from the cylinder flows, a supercharger provided in the intake passage to supercharge the intake air, and a supercharger provided in the intake passage. A bypass passage that bypasses the supercharger and an EGR device that recirculates EGR gas, which is a part of the exhaust gas in the exhaust passage, to a portion of the intake passage upstream of the supercharger. It is an engine control device
Combustion suppression means that can suppress the progress of combustion of the air-fuel mixture in the cylinder,
A determination means for determining whether or not the EGR rate, which is the ratio of the amount of EGR gas in the cylinder to the total amount of gas in the cylinder, is less than a predetermined reference EGR rate.
A control means for controlling the supercharger and the combustion suppression means is provided.
When the determination means determines that the EGR rate is less than the reference EGR rate at the start of supercharging for driving the supercharger in the stopped state, the control means of the air-fuel mixture in the cylinder. When the combustion slowdown control for controlling the combustion suppression means is performed so that the progress of combustion is suppressed and the determination means determines that the EGR rate is equal to or higher than the reference EGR rate, the combustion slowdown is performed. An engine control device characterized in that the execution of control is prohibited. In the engine control device according to claim 1,
A fuel injection means that injects fuel into the cylinder and functions as the combustion suppression means is provided.
When the combustion slowdown control is executed, the control means controls the fuel injection means so that the execution time of fuel injection to the cylinder is delayed as compared with the non-execution time of the combustion slowdown control. Engine control device. In the engine control device according to claim 2.
The fuel injection means performs a first fuel injection and a second fuel injection whose injection timing is later than that of the first fuel injection.
An engine control device, characterized in that, when the combustion slowdown control is executed, the control means controls the fuel injection means so that the second fuel injection is performed in the latter half of the compression stroke. In the engine control device according to claim 3,
When the combustion slowdown control is executed, the control means increases the total amount of fuel supplied to the cylinder in one combustion cycle as the amount of air introduced into the cylinder is large and the amount of EGR gas introduced into the cylinder is small. An engine control device, characterized in that the fuel injection means is controlled so that the ratio of the amount of fuel supplied to the cylinder by the second fuel injection with respect to the second fuel injection is large. In the engine control device according to claim 3 or 4.
An engine control device, characterized in that, when the combustion slowdown control is executed, the control means executes both the first fuel injection and the second fuel injection in a compression stroke. In the engine control device according to any one of claims 1 to 5.
It has an air filling amount changing means that can change the air filling amount, which is the amount of air introduced into the cylinder.
The control means fills the air when the EGR rate insufficient condition is satisfied, in which the EGR rate is larger than the preset target EGR rate and the difference between the EGR rate and the target EGR rate is equal to or greater than a predetermined determination value. An engine control device for controlling the air filling amount changing means so that the amount is reduced. In the engine control device according to claim 6,
The control means changes the air filling amount so that the air filling amount decreases when the EGR rate insufficient condition is satisfied and when the temperature of the intake air introduced into the cylinder is equal to or higher than a predetermined determined intake air temperature. An engine control device characterized by controlling means. In the engine control device according to any one of claims 1 to 7.
A purification device provided in the exhaust passage and capable of purifying NOx in the exhaust gas is provided.
The EGR device is an engine control device characterized in that exhaust gas in the exhaust passage on the downstream side of the purification device is returned to the intake passage. In the engine control device according to any one of claims 1 to 8.
Equipped with an ignition means that ignites the air-fuel mixture in the cylinder
The control means is characterized in that the ignition means is controlled so that a part of the air-fuel mixture in the cylinder is spark-ignited and burned by ignition from the ignition means and the remaining air-fuel mixture is burned by self-ignition. Engine control device.