JP7205356B2 - Engine combustion control device - Google Patents

Description

The present invention relates to a combustion control device for an engine having a plurality of cylinders and capable of executing internal EGR operation in each cylinder.

In an engine having a plurality of cylinders, it is important to suppress variations in the air-fuel ratio among the cylinders from the viewpoint of stabilizing the output and improving the fuel efficiency. Patent Document 1 discloses a method of determining the air-fuel ratio of each cylinder and controlling the air-fuel ratio of each cylinder based on the detected value of an air-fuel ratio sensor arranged in an exhaust collecting portion. However, this method has the drawback that it is difficult to determine the air-fuel ratio of each cylinder during low-load operation when the flow rate of exhaust gas is low, and accurate air-fuel ratio control cannot be performed.

By the way, in order to adjust the in-cylinder temperature of the cylinder to a desired temperature, internal EGR is widely used in which high-temperature burned gas remains in the cylinder. In general, internal EGR is achieved by setting valve overlap to open both the intake valve and the exhaust valve. For example, in an engine that performs compression self-ignition combustion, internal EGR can raise the temperature of the air-fuel mixture in the cylinder to a temperature at which self-ignition is possible. In order to control the air-fuel ratio of the cylinders, and thus to control the variation in the air-fuel ratio among the cylinders, it is necessary to consider the amount of intake air (fresh air) re-intake into the cylinders from the intake port by the internal EGR.

JP-A-8-232729

Engines are often required to be compact from the viewpoint of mountability in vehicles, and there are cases where it is required to set the paths of the intake system and the exhaust system as short as possible. For example, the intake system has an independent intake passage that connects the surge tank and the intake port of each cylinder, and there are cases where this independent intake passage must be set short. In this case, during the period of valve overlap for internal EGR, the intake air that is once blown out from the intake port of one cylinder due to the intake-exhaust differential pressure reaches the surge tank through the independent intake passage, and then goes to the cylinder instead of the cylinder itself. A phenomenon of re-intake into other cylinders may occur. This phenomenon causes a difference in the reintake amount of intake air among the cylinders, resulting in variations in the air-fuel ratio among the cylinders.

SUMMARY OF THE INVENTION An object of the present invention is to provide an engine combustion control device capable of suppressing variations in air-fuel ratio among a plurality of cylinders in an engine in which internal EGR is performed.

An engine combustion control device according to one aspect of the present invention connects a plurality of cylinders, a surge tank disposed in an intake path to the plurality of cylinders, and the surge tank and each intake port of the plurality of cylinders. an independent intake passage, an intake valve and an exhaust valve for opening and closing each intake port and each exhaust port of the plurality of cylinders, and fuel arranged for each of the plurality of cylinders to supply fuel to each cylinder An internal EGR mechanism that realizes internal EGR by setting a valve overlap in which the injection valve, the intake valve and the exhaust valve are both opened, and the fuel injection amount of each of the fuel injection valves are controlled by the operation of the engine. a control unit that controls according to the state, and a first combustion mode in which partial compression ignition combustion combining SI combustion and CI combustion is performed in an environment that substantially matches the stoichiometric air-fuel ratio, and in a lean environment. wherein the independent intake passage allows the intake air blown out from the intake port to reach the surge tank during the valve overlap period. In both the first combustion mode and the second combustion mode, the control unit sets the target fuel injection amount for each cylinder, which is determined according to the operating state of the engine , to each cylinder. is corrected based on a re-intake correction amount set for each cylinder according to the re-intake amount of intake air from the intake port due to internal EGR in the first combustion mode. It is characterized in that the re-intake correction amount is reduced as the engine speed increases .

According to this combustion control device, even if the re-intake amount of intake air due to internal EGR varies among a plurality of cylinders for some reason, the target fuel injection amount is corrected for each cylinder by the re-intake correction amount. Further, the re-intake correction amount is corrected according to the engine speed. Engine speed is proportional to the time that valve overlap is currently occurring. In other words, whether the engine speed is high or low leads to the length of the valve overlap time. Therefore, the engine speed affects the amount of burned gas returned to the cylinder by internal EGR, and further affects the amount of intake air blown out from the intake port. If the blown intake air amount fluctuates, the reinhalation mode may also fluctuate. Therefore, instead of setting the reinhalation correction amount fixedly, By correcting according to the engine speed, the air-fuel ratio can be adjusted more accurately for each cylinder. As a result, it is possible to accurately match the air-fuel ratios among the plurality of cylinders by reflecting the operating conditions.

In the engine combustion control device described above, it is preferable that the independent intake passage is set to a length that allows intake air blown out from the intake port to reach the surge tank during the valve overlap period.

According to this combustion control device, the independent intake passage is set to have a length that allows the intake air discharged from the intake port by the internal EGR to reach the surge tank. For this reason, while the vehicle-mountability of the engine can be improved, the intake air once discharged from the intake port of one cylinder is not re-inhaled into its own cylinder, but is re-inhaled into the other cylinders through the surge tank. phenomenon can occur. In other words, the amount of reintake of the intake air may differ among the cylinders. Even if such a difference in the re-intake amount occurs, the target fuel injection amount is corrected for each cylinder by the re-intake correction amount, so variations in the air-fuel ratio between cylinders can be suppressed.

In the engine combustion control device described above, it is preferable that the control unit corrects the re-intake correction amount so as to decrease as the engine speed increases.

As the engine speed increases, the valve overlap time shortens and the amount of burned gas returned to the cylinder decreases. Along with this, the amount of intake air blown out from the intake port also decreases, so the variation in the amount of reintake of intake air among the cylinders tends to decrease. According to the above combustion control device, the re-suction correction amount is reduced as the engine speed increases, so that variations in the air-fuel ratio among the cylinders can be suppressed accurately.

In the engine combustion control device described above, it is preferable that the control unit corrects the re-intake correction amount so as to decrease as the intake pressure increases.

When the intake pressure is high, as the intake/exhaust differential pressure becomes smaller, the amount of intake air blown out from the intake port decreases, and the re-intake amount naturally decreases. Therefore, since the variation in the re-intake amount among the cylinders is also reduced, the target fuel injection amount is corrected in accordance with the actual situation by correcting so that the re-intake correction amount is decreased as the intake pressure increases. can do.

In the above engine combustion control device, among the plurality of cylinders, a first cylinder having a predetermined amount of re-intake of intake air from the intake port due to the internal EGR, and a first cylinder having a predetermined amount of re-intake of intake air from the intake port. The control unit adjusts the air-fuel ratio of the first cylinder to the lean side and the air-fuel ratio of the second cylinder to the rich side. It is desirable to correct the target fuel injection amount.

According to this combustion control device, the first cylinder, in which the amount of re-intake of the intake air from the intake port is relatively small, is leaned, that is, the fuel injection amount is reduced, while the second cylinder, in which the amount of fuel injection is relatively large, is reduced. As for the cylinder, the target fuel injection amount is corrected so as to increase the fuel injection amount on the rich side. Therefore, the fuel injection amount is corrected for each cylinder according to the re-intake characteristic inherent to the engine, and the air-fuel ratio among the cylinders can be adjusted to a constant value (eg, λ=1).

In the engine combustion control device described above, the engine is a four-cylinder engine in which four cylinders are arranged in a row, the independent intake passages are arranged in a row in the direction in which the cylinders are arranged, and the surge tank is longitudinal in the direction in which the cylinders are arranged. and the surge tank is connected to an upstream intake passage that introduces intake air into a central region in the longitudinal direction of the surge tank. The target fuel injection is performed such that the air-fuel ratios of the end cylinders positioned at both ends of the cylinders are leaner, and the air-fuel ratios of the two central cylinders sandwiched between the end cylinders are richer. It is desirable to correct the amount.

In the intake system configuration described above, among the four cylinders, the end cylinders have a relatively small amount of reintake of the intake air from the intake ports, and the center side cylinder has a relatively large amount of reintake. In other words, there is a tendency that part of the intake air once discharged from the end-side cylinder intake ports by the internal EGR is re-intaken into the intake ports of the center-side cylinders. According to the above combustion control device, the end cylinders with relatively small intake air reintake amount are leaned (reduced fuel injection amount), while the central cylinders with relatively large amount are rich side (reduced fuel injection amount). The target fuel injection amount is corrected by increasing the fuel injection amount. Therefore, the fuel injection amount is corrected for each cylinder according to the characteristic re-intake characteristics based on the configuration of the intake system described above, and the air-fuel ratio among the four cylinders can be adjusted to a constant value (eg, λ=1).

In the engine combustion control device described above, the control unit controls at least an amount of valve overlap at which both the intake valve and the exhaust valve of the cylinder are in an open state, an intake/exhaust differential pressure, and closing of the exhaust valve. It is desirable to set the re-intake correction amount based on the re-intake amount calculated using a polynomial model having timing and engine speed as elements.

According to this combustion control device, the control unit can easily and accurately set the reintake correction amount using the polynomial model.

Advantageous Effects of Invention According to the present invention, it is possible to provide an engine combustion control device capable of suppressing variations in air-fuel ratio among a plurality of cylinders in an engine in which internal EGR is performed.

FIG. 1 is a system diagram schematically showing the overall configuration of a compression ignition engine to which an engine combustion control device according to the present invention is applied. FIG. 2 is a perspective view showing the appearance of the engine body. FIG. 3 is a longitudinal sectional view of the engine body showing the structure of the intake passage. FIG. 4 is a vertical cross-sectional view around a surge tank arranged in an intake passage. FIG. 5 is a partially broken perspective view of an exhaust manifold. FIG. 6 is a block diagram showing an engine control system. FIGS. 7A to 7C are engine operation maps in warm, semi-warm, and cold conditions, respectively. FIG. 8 is a graph showing the waveform of the heat release rate during SPCCI combustion. 9A is a graph showing the valve overlap period during λ=1 combustion, FIG. 9B is a graph showing the valve overlap period during lean combustion, and FIG. 9C is a graph showing the return amount of burned gas and the valve overlap time. is a graph showing the relationship between FIG. 10 is a schematic diagram for explaining the state of the return of exhaust gas to the cylinder, the return of intake air, and re-intake due to the execution of internal EGR. FIGS. 11A to 11D are graphs showing correction tendencies of the fuel injection amount of each cylinder in the operating region of λ=1. FIGS. 12A to 12D are graphs showing correction trends of the fuel injection amount of each cylinder in the operating region of λ>1. FIGS. 13A to 13C are graphs respectively showing the variables dependent on the operating state and the tendency of correction of the specific re-inhalation correction amount. FIGS. 14A to 14C are graphs respectively showing the variables dependent on the operating state and the tendency of correction of the inherent re-inhalation correction amount. FIG. 15 is a schematic diagram showing the process of obtaining the internal EGR amount predicted value. FIG. 16 is an explanatory diagram showing the procedure for correcting the fuel injection amount. FIG. 17 is a flowchart showing a specific example of combustion control according to this embodiment.

BEST MODE FOR CARRYING OUT THE INVENTION Hereinafter, embodiments of the present invention will be described with reference to the drawings. In this embodiment, an example in which an engine combustion control device according to the present invention is applied to a compression ignition engine will be described. Of course, the compression ignition engine is an engine capable of performing internal EGR operation in each cylinder.

[Engine system]
First, an engine system including the compression ignition engine will be described. FIG. 1 is a diagram showing the overall configuration of an engine system according to this embodiment. The engine system comprises an engine body 1 consisting of a 4-cycle 4-cylinder gasoline direct injection engine, an intake passage 30 through which intake air introduced into the engine body 1 flows, and an exhaust passage through which exhaust gas discharged from the engine body 1 flows. 40 and an EGR device 50 that recirculates part of the exhaust gas flowing through the exhaust passage 40 to the intake passage 30 .

The engine body 1 includes a cylinder block 3 , a cylinder head 4 and pistons 5 . The cylinder block 3 has a cylinder liner that forms four cylinders 2 (plurality of cylinders) arranged in a line. Note that only one cylinder 2 is shown in FIG. The cylinder head 4 is attached to the upper surface of the cylinder block 3 and closes the upper opening of the cylinder 2 . A piston 5 is housed in each cylinder 2 so as to be reciprocally slidable, and is connected to a crankshaft 7 via a connecting rod 8 . As the piston 5 reciprocates, the crankshaft 7 rotates about its central axis.

A combustion chamber 6 is formed above the piston 5 . A fuel containing gasoline as a main component is supplied to the combustion chamber 6 by injection from an injector 15, which will be described later. A mixture of the supplied fuel and air is combusted in the combustion chamber 6, and the piston 5, which is pushed down by the expansion force of the combustion, reciprocates vertically. The geometric compression ratio of the cylinder 2, that is, the ratio between the volume of the combustion chamber 6 when the piston 5 is at top dead center and the volume of the combustion chamber 6 when the piston 5 is at bottom dead center, is determined by SPCCI combustion, which will be described later. is set to a high compression ratio of 13 or more and 30 or less, preferably 14 or more and 18 or less.

A crank angle sensor SN1 and a water temperature sensor SN2 are attached to the cylinder block 3 . The crank angle sensor SN1 is arranged to detect the rotation angle (crank angle) of the crankshaft 7 and the rotation speed of the crankshaft 7, that is, the engine speed. The water temperature sensor SN2 detects the temperature of cooling water flowing inside the cylinder block 3 and the cylinder head 4, that is, the engine water temperature.

An intake port 9 and an exhaust port 10 communicating with the combustion chamber 6 are formed in the cylinder head 4 . The bottom surface of the cylinder head 4 serves as the ceiling surface of the combustion chamber 6 . An intake side opening that is the downstream end of the intake port 9 and an exhaust side opening that is the upstream end of the exhaust port 10 are formed in the ceiling surface of the combustion chamber. 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 are assembled to the cylinder head 4 . Although not shown, the engine body 1 has a 4-valve format consisting of 2 intake valves and 2 exhaust valves, and two intake ports 9 and two exhaust ports 10 are provided for each cylinder 2. In addition, two intake valves 11 and two exhaust valves 12 are also provided.

The cylinder head 4 is provided with an intake-side valve mechanism 13 and an exhaust-side valve mechanism 14 including camshafts. The intake valve 11 and the exhaust valve 12 are driven to open and close by these valve mechanisms 13 and 14 in conjunction with the rotation of the crankshaft 7 . The intake valve mechanism 13 incorporates an intake VVT 13 a capable of changing at least the opening timing of the intake valve 11 . Similarly, the exhaust side valve mechanism 14 incorporates an exhaust VVT 14a capable of changing at least the closing timing of the exhaust valve 12 . By controlling the intake VVT 13a and the exhaust VVT 14a, it is possible to set a valve overlap in which both the intake valve 11 and the exhaust valve 12 are opened across the exhaust top dead center (internal EGR mechanism). By setting this valve overlap, internal EGR is realized in which high-temperature burned gas remains in the combustion chamber 6 . Further, it is possible to adjust the internal EGR amount (remaining amount of burned gas) by adjusting the amount of valve overlap, which is the period during which valve overlap is performed.

Injectors 15 (fuel injection valves) and spark plugs 16 are also arranged in the cylinder head 4 for each of the four cylinders 2 . The injector 15 injects (supplies) fuel into the cylinder 2 . As the injector 15, it is possible to use a multi-hole type injector that has a plurality of injection holes at its tip and is capable of radially injecting fuel from these injection holes. The injector 15 is arranged such that its tip is exposed in the combustion chamber 6 and faces the radial center of the crown surface of the piston 5 .

The spark plug 16 is arranged at a position slightly shifted toward the intake side with respect to the injector 15 , and is arranged at a position where the tip electrode faces the inside of the cylinder 2 . The spark plug 16 is a forced ignition source that ignites a mixture of fuel and air formed in the cylinder 2 (combustion chamber 6).

The cylinder head 4 is provided with an in-cylinder pressure sensor SN3, an intake cam angle sensor SN12, and an exhaust cam angle sensor SN13 as sensing elements. An in-cylinder pressure sensor SN3 detects the pressure in the combustion chamber 6 . An intake cam angle sensor SN12 detects the rotational position of the camshaft of the intake side valve mechanism 13 . The exhaust cam angle sensor SN13 detects the rotational position of the camshaft of the exhaust-side valve train 14 .

The intake passage 30 is connected to one side surface of the cylinder head 4 so as to communicate with the outside air and the intake port 9 . Air (fresh air) taken 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 . An air cleaner 31 , a throttle valve 32 , a supercharger 33 , an electromagnetic clutch 34 , an intercooler 35 , a surge tank 36 and an independent intake passage 37 are arranged in the intake passage 30 in this order from the upstream side.

The air cleaner 31 cleans the intake air by removing foreign substances in the intake air. The throttle valve 32 opens and closes the intake passage 30 in conjunction with the depression of the accelerator 17 to adjust the flow rate of intake air in the intake passage 30 . The supercharger 33 compresses the intake air and sends the intake air to the downstream side of the intake passage 30 . The supercharger 33 is a supercharger mechanically linked to the engine body 1 , and is switched between engagement with and disengagement from the engine body 1 by an electromagnetic clutch 34 . When the electromagnetic clutch 34 is engaged, driving force is transmitted from the engine body 1 to the supercharger 33, and supercharging by the supercharger 33 is performed. The intercooler 35 cools the intake air compressed by the supercharger 33 .

The surge tank 36 is a tank that is arranged in an intake path to the four cylinders 2 and provides a space for evenly distributing the intake air to the four cylinders 2 . The independent intake passage 37 is arranged downstream of the surge tank 36 and is a passage that independently connects the surge tank 36 and the intake ports 9 of the four cylinders 2 . After the intake air supplied from the air cleaner 31 side flows into the air cleaner 31 , it is supplied to the intake port 9 of each cylinder 2 through each independent intake passage 37 . One of the two intake ports 9 provided corresponding to each cylinder 2 is provided with a swirl valve 37A. By adjusting the opening degree of the swirl valve 37A, the strength of the swirl flow swirling around the central axis of the combustion chamber 6 can be adjusted. The intake port 9 of this embodiment is a tumble port capable of forming a tumble flow. Therefore, the swirl flow formed when the swirl valve 37A is closed becomes an oblique swirl flow mixed with the tumble flow.

Each portion of the intake passage 30 includes an air flow sensor SN4 for detecting the flow rate of intake air, first and second intake air temperature sensors SN5 and SN7 for detecting the temperature of the intake air, and first and second intake air temperature sensors for detecting the pressure of the intake air. Air pressure sensors SN6 and SN8 are provided. An airflow sensor SN4 and a first intake air temperature sensor SN5 are arranged in a portion between the air cleaner 31 and the throttle valve 32 in the intake passage 30, and detect the flow rate and temperature of intake air passing through that portion, respectively. The first intake pressure sensor SN6 is provided in a portion of the intake passage 30 between the throttle valve 32 and the supercharger 33 (downstream of a connection port of the EGR passage 51, which will be described later). Detect pressure. The second intake air temperature sensor SN7 is provided in a portion between the supercharger 33 and the intercooler 35 in the intake passage 30, and detects the temperature of the intake air passing through that portion. A second intake pressure sensor SN8 is provided in the surge tank 36 and detects the pressure of intake air in the surge tank 36 .

The intake passage 30 is provided with a bypass passage 38 for bypassing the supercharger 33 and sending intake air to the combustion chamber 6 . The bypass passage 38 connects the surge tank 36 and the vicinity of the downstream end of the EGR passage 51 to be described later. A bypass valve 39 capable of opening and closing the bypass passage 38 is provided in the bypass passage 38 .

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 . Burned gas (exhaust gas) 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 arranged in the exhaust passage 40 . The catalytic converter 41 includes a three-way catalyst 41a for purifying harmful components (HC, CO, NOx) contained in the exhaust gas flowing through the exhaust passage 40, and particulate matter (PM) contained in the exhaust gas. A GPF (gasoline particulate filter) 41b for collecting is incorporated. Note that another catalytic converter containing an appropriate catalyst such as a three-way catalyst or a NOx catalyst may be added downstream of the catalytic converter 41 .

A linear _O2 sensor SN9 for detecting the concentration of oxygen contained in the exhaust gas and an exhaust temperature sensor SN10 for measuring the temperature of the exhaust gas are provided in the exhaust passage 40 upstream of the catalytic converter 41. ing. The linear _O2 sensor SN9 is a type of sensor whose output value linearly changes according to the density of oxygen concentration. Based on the output value of the linear _O2 sensor SN9, it is possible to estimate the air-fuel ratio of the air-fuel mixture. The measured value of the exhaust temperature sensor SN10 is used for calculating the predicted value of the internal EGR amount.

The EGR device 50 includes an EGR passage 51 connecting 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 downstream of the catalytic converter 41 and a portion of the intake passage 30 between the throttle valve 32 and the supercharger 33 . The EGR cooler 52 cools the exhaust gas (external EGR) recirculated from the exhaust passage 40 to the intake passage 30 through the EGR passage 51 by heat exchange. The EGR valve 53 is provided in the EGR passage 51 on the downstream side of the EGR cooler 52 so as to be openable and closable, 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 SN11 for detecting the difference between the pressure on the upstream side of the EGR valve 53 and the pressure on the downstream side thereof.

The accelerator 17 is provided with an accelerator opening sensor SN14 for detecting the accelerator opening. The accelerator opening sensor SN14 is a sensor that detects the degree of depression of the accelerator 17, and is also a sensor that detects the driver's intention to accelerate or decelerate.

[Structural Features of Intake Passage and Exhaust Passage]
Next, structural features of the intake passage 30 and the exhaust passage 40 described above will be described. 2 is a perspective view showing the exterior of the engine body 1, FIG. 3 is a longitudinal sectional view of the engine body 1 showing the structure of the intake passage 30, and FIG. 4 is a surge tank 36 disposed in the intake passage 30. Fig. 2 is a longitudinal sectional view in the left-right direction of the engine body 1 showing the surroundings; Although these figures are labeled with up/down, front/rear, and left/right directions, this is for convenience of explanation and is not meant to limit the actual directions.

The intake passage 30 is attached to the front surface 101 side of the engine body 1 . Although not shown in FIGS. 2 and 3, the exhaust passage 40 is attached to the rear surface side of the engine body 1. As shown in FIG. 1, the intake passage 30 includes a first passage 301, a second passage 302 and a third passage 303. As shown in FIG. The first passage 301 is an intake passage that connects between the air cleaner 31 and the supercharger 33 . A downstream end of the EGR passage 51 joins the first passage 301 downstream of the throttle valve 32 . The second passage 302 is an intake passage connecting between the supercharger 33 and the intercooler 35 . A third passage 303 is an intake passage connecting between the intercooler 35 and the surge tank 36 .

The supercharger 33 is attached to the front surface 101 near the upper right side of the engine body 1 . The first passage 301 extends to project rightward from the supercharger 33 and guides the air cleaned by the air cleaner 31 to the supercharger 33 . The intercooler 35 is attached to the front surface 101 near the lower portion of the engine body 1 . The second passage 302 extends vertically in the center of the engine body 1 in the left-right direction, and guides air (and external EGR) that has passed through the supercharger 33 to the intercooler 35 . Intercooler 35 includes a water-cooled core 351 and a cooler housing 352 that supports core 351 . A cooler housing 352 connects the downstream end of the second passage 302 and the upstream end of the third passage 303 .

The surge tank 36 is arranged near the upper portion of the engine body 1 so as to face the upstream end of the intake port 9 and be adjacent to the rear surface of the supercharger 33, as shown in FIG. The third passage 303 extends vertically in the center of the engine body 1 in the left-right direction, and guides the air that has passed through the intercooler 35 and is cooled to the surge tank 36 . The bypass passage 38 extends in the left-right direction near the upper portion of the engine body 1 .

The independent intake passage 37 extends in the longitudinal direction so as to connect the surge tank 36 and the upstream end of the intake port 9 . FIG. 3 shows the independent intake passage 37 in which the swirl valve 37V is arranged. The independent intake passage 37 of the present embodiment has a short passage length, and is set to a passage length shorter than the width of the internal space 36R of the surge tank 36 in the front-rear direction. Specifically, the channel length of the independent intake passage 37 is set to be approximately half the width of the internal space 36R in the front-rear direction and slightly longer than the valve element length of the swirl valve 37V. That is, when the swirl valve 37V is fully opened, one end of the valve element enters the intake port 9 and the other end enters the surge tank 36 . With such a short independent intake passage 37, the engine system can be made compact, and the mountability on the vehicle is improved. However, the passage length of the independent intake passage 37 is also a length that allows the intake air blown out from the intake port 9 to reach the surge tank 36 during the valve overlap period for executing internal EGR. Problems caused by this will be described later with reference to FIG.

In FIGS. 2 and 3, the flow of intake air in the intake passage 30 is indicated by arrows F1 to F5. Intake air is introduced into the first passage 301 through the air cleaner 31 as indicated by an arrow F1. Next, the intake air passes through the scroll portion of the compressor of the supercharger 33 and is blown out to the second passage 302 . The intake air then flows from top to bottom through the second passage 302 (arrow F2), enters the cooler housing 352 of the intercooler 35, and passes through the core 351 (arrow F3). After that, the intake air flows through the third passage 303 from bottom to top (arrow F4) and reaches the internal space 36R of the surge tank 36 . Thereafter, the intake air changes its flow direction to the front-rear direction, passes through each independent intake passage 37, and flows into the intake port 9 (arrow F5).

The arrangement relationship between the surge tank 36 and each independent intake passage 37 will be described with reference to FIG. A housing that forms the surge tank 36 is integrally formed with a housing that forms the third passage 303 (upstream intake passage). The housing of the third passage 303 includes a collecting portion 304 that collects the intake air that has passed through the intercooler 35 and an introduction portion 305 that continues downstream from the collecting portion 304 and is connected to the surge tank 36 .

The surge tank 36 has an internal space 36R (flow path space) elongated in the left-right direction. The longitudinal direction of the internal space 36R is the arrangement direction (horizontal direction) in which the four cylinders 2 are arranged in a line. A downstream end of the introduction portion 305 (the third passage 303 ) communicates with the internal space 36</b>R through an introduction port 362 opened in the bottom wall 361 of the surge tank 36 . The inlet 362 is arranged in the longitudinal central region of the bottom wall 361 . That is, the third passage 303 has a structure (hereinafter referred to as "center intake structure") that introduces intake air into the central region in the longitudinal direction of the surge tank 36, as indicated by the arrow F4.

The independent intake passage 37 extends rearward from an opening provided in a rear wall 363 of the surge tank 36 that is perpendicular to the bottom wall 361 . The rear wall 363 is a wall elongated in the left-right direction. As described above, each cylinder 2 is of the 4-valve type, so each cylinder 2 is provided with two independent intake passages 37 . In FIG. 4, the four cylinders 2 are represented as #1 to #4 cylinders 2, the two independent intake passages corresponding to #1 cylinder 2 are designated as "37_#1", and the two independent intake passages corresponding to #2 cylinder 2 are designated as "37_#1". The intake passage is indicated as "37_#2". A total of eight independent intake passages 37_#1 to 37_#4 are arranged in a line in the horizontal direction, which is the direction in which the four cylinders 2 are arranged. The independent intake passages 37_#1 and 37_#4 of #1 and #4 cylinders 2 are relatively far from the introduction port 362, and the independent intake passages 37_#2 and 37_#3 of #2 and #3 cylinders 2 are relatively far. It is in a positional relationship that it is close to the target.

Next, structural features of the exhaust passage 40 will be described. As shown in FIG. 5 , the exhaust passage 40 has an exhaust manifold 42 (exhaust collecting portion) connected to the exhaust port 10 of the engine body 1 . The exhaust manifold 42 collects exhaust gases discharged from the exhaust ports 10 of the #1 to #4 cylinders 2 . The exhaust collected in the exhaust manifold 42 is guided downstream through an exhaust pipe 43 (downstream exhaust passage) that continues to the catalytic converter 41 .

The exhaust manifold 42 joins the independent exhaust pipes 44_#1 to 44_#4 (exhaust paths) corresponding to the #1 to #4 cylinders 2, respectively, and the exhaust guided by the independent exhaust pipes 44_#1 to 44_#4. and a fastening flange 46 to be attached to the rear surface of the engine body 1 . As is clear from FIG. 5, the independent exhaust pipe 44_#1 corresponding to the #1 cylinder 2 (one end cylinder) located at the right end (one end) of the #1 to #4 cylinders 2 is connected to the exhaust port 10. to the exhaust pipe 43 on the downstream side is the shortest. As for the independent exhaust pipes 44_#2, 44_#3, and 44_#4 corresponding to the #2 to #4 cylinders 2, the cylinders positioned on the left (other end side) of #2, #3, and #4 2, the path from the exhaust port 10 to the exhaust pipe 43 is long. In other words, the independent exhaust pipe 44_#1 goes straight toward the exhaust pipe 43 with the shortest flow path length, while the independent exhaust pipe 44_ with respect to the exhaust pipe 43 moves toward #2, #3, and #4 toward the left. This is a structure in which the channel lengths of #2 to 44_#4 are lengthened (hereinafter referred to as "exhaust structure unevenly distributed at the end").

[Control configuration]
FIG. 6 is a block diagram showing the control configuration of the engine system. The engine system of this embodiment is centrally controlled by a processor 60 (engine combustion control device). The processor 60 is a microprocessor comprising a CPU, ROM, RAM and the like.

Detection signals from various sensors are input to the processor 60 . The processor 60 uses the aforementioned crank angle sensor SN1, water temperature sensor SN2, in-cylinder pressure sensor SN3, airflow sensor SN4, first and second intake air temperature sensors SN5 and SN7, first and second intake air pressure sensors SN6 and SN8, linear O ₂ sensor SN9, exhaust temperature sensor SN10, differential pressure sensor SN11, intake cam angle sensor SN12, exhaust cam angle sensor SN13, accelerator opening sensor SN14 and atmospheric pressure sensor SN15. The atmospheric pressure sensor SN15 is a sensor that measures the atmospheric pressure of the running environment, and is used exclusively to detect the running altitude. Information detected by these sensors SN1 to SN15, that is, crank angle, engine speed, engine water temperature, cylinder pressure, intake flow rate, intake air temperature, intake pressure, exhaust gas oxygen concentration, exhaust gas temperature, EGR valve Information such as the differential pressure across the engine 53, intake/exhaust cam angles, accelerator opening, atmospheric pressure, etc. are sequentially input to the processor 60. FIG.

The processor 60 controls each part of the engine while executing various judgments and calculations based on the input information from each sensor. That is, the processor 60 is electrically connected to the intake VVT 13a, the exhaust VVT 14a, the injector 15, the spark plug 16, the throttle valve 32, the electromagnetic clutch 34, the bypass valve 39, the EGR valve 53, and the like. Based on this, control signals are output to each of these devices.

The ECU 100 operates to functionally include a fuel injection control section 61 (control unit), an ignition control section 67 and a storage section 68 by executing a predetermined program. The fuel injection control unit 61 controls the amount of fuel injected by each injector 15 of cylinders #1 to #4 according to the operating state of the engine. The ignition control section 67 controls the ignition operation of the spark plug 16 . The storage unit 68 stores various data, set values, arithmetic expressions, and the like. In the present embodiment, the storage unit 68 stores a re-intake correction amount specific to internal EGR, which will be described later, a polynomial model for estimating the internal EGR amount, and the like.

The fuel injection control unit 61 functionally includes an operating state determination unit 62, an injection setting unit 63, an EGR prediction unit 64, a first correction unit 65, and a second correction unit 66 by executing a predetermined program. works like

The operating state determination unit 62 determines the operating state of the engine body 1 from the engine speed based on the detection value of the crank angle sensor SN1 and the engine load based on the opening information of the accelerator opening sensor SN14. This determination result is used to determine which area of the predetermined operation map the current operating area is.

The injection setting unit 63 sets the fuel injection amount and injection pattern from the injector 15 according to various conditions. The injection setting unit 63 first sets a target fuel injection amount and an injection pattern according to the accelerator depression amount (engine operating state) detected by the accelerator opening sensor SN14. Further, the injection setting unit 63 refers to the primary correction amount derived by the first correction unit 65 and the secondary correction amount derived by the second correction unit 66 according to the execution state of internal EGR, that is, at least Correct the target fuel injection amount.

The EGR prediction unit 64 estimates the return amount (internal EGR amount) of the burned gas from the exhaust passage 40 side into the cylinder 2 by the internal EGR executed by setting the valve overlap period of the intake valve 11 and the exhaust valve 12. Perform processing to obtain the predicted value. The return amount of the burned gas is also the amount of the intake air blown out from the cylinder 2 toward the intake passage 30 by the return of the burned gas, and the amount of the blown intake air re-inhaled from the intake port 9. The EGR prediction unit 64 applies the polynomial model stored in the storage unit 68 to obtain the internal EGR amount.

The first correction unit 65 primarily corrects the target fuel injection amount set by the injection setting unit 63 according to the operating state. The primary correction amount by the first correction unit 65 is a specific correction amount determined by the engine structure, particularly the structures of the intake passage 30 and the exhaust passage 40 . In this embodiment, the intake passage 30 has a center intake structure (FIG. 4), and the exhaust passage 40 has an end unevenly distributed exhaust structure (FIG. 5). Due to these structures, the re-intake amount of intake air from the intake port 9 or the return amount of burned gas from the exhaust port 10 in #1 to #4 cylinders 2 due to internal EGR may differ depending on the operating conditions. . For example, in an operating region where the amount of valve overlap is relatively small and the amount of internal EGR is relatively small, the center intake structure affects the intake air reintake amount for each cylinder 2, and the air-fuel ratio between #1 to #4 cylinders 2 is variation. On the other hand, in the operating region where the amount of valve overlap is relatively large and the amount of internal EGR is relatively large, the end unevenly distributed exhaust structure affects the return amount of burned gas for each cylinder 2, and the #1 to #4 cylinders 2 causes variations in the gas air-fuel ratio (described later with reference to FIG. 10).

In an operating region where the internal EGR amount is relatively small (an operating region of SPCCI_λ=1, which will be described later), the first correction unit 65 adjusts the #1 ˜#4 The target fuel injection amount is primarily corrected using the re-intake correction amount set for each cylinder 2 as the primary correction amount. On the other hand, in an operating region where the internal EGR amount is relatively large (an operating region where SPCCI_λ>1, which will be described later), the first correction unit 65 adjusts the amount of burned gas returned from each exhaust port 10 of cylinders #1 to #4. Accordingly, the target fuel injection amount is primarily corrected using the return correction amount set for each of #1 to #4 cylinders 2 as the primary correction amount. The re-suction correction amount and the return correction amount are stored in the storage unit 68 for each cylinder #1 to #4 in association with the internal EGR amount and the correction amount. Alternatively, the return correction amount is read and the primary correction processing is performed.

The second corrector 66 secondarily corrects at least the unique reinhalation correction amount set by the first corrector 65 according to variables based on the operating state and environmental conditions. Examples of the variables include intake pressure, valve overlap amount, closing timing of the exhaust valve 12, engine speed, atmospheric pressure, and exhaust gas temperature. As the variables listed above change, the amount of reinspired air from each intake port 9 may also change. Therefore, if only the re-intake correction amount unique to the engine structure is applied, there may be a case where the air-fuel ratios of #1 to #4 cylinders 2 cannot be matched accurately. In view of this point, the second correction unit 66 performs secondary correction to increase or decrease the unique reinhalation correction amount set by the first correction unit 65 according to the situation of the variables. The secondary correction amount by the second correction unit 665 is predetermined according to the degree of deviation of the variable from the reference value.

[Driving map]
FIGS. 7A to 7C are operation maps for explaining differences in control according to the progress of engine warm-up and the rotation speed/load of the engine. In the present embodiment, a first operation map Q1 (FIG. 7A) is used when the engine is warmed up, and a second operation map is used when the engine is halfway warmed up. Q2 (FIG. 7(B)) and a third operation map Q3 (FIG. 7(C)) used when the engine is not yet warmed up and are cold are prepared. The warm first operation map Q1 includes first, second, third, fourth, and fifth operation regions A1, A2, A3, A4, and A5 with different combustion modes. The second operating map Q2 for hour includes sixth, seventh, eighth, and ninth operating regions B1, B2, B3, and B4 with different combustion modes. The cold third operation map Q3 consists of one of the tenth operation regions C1.

<During warmth>
In the first operation map Q1, the first area A1 is a low/medium speed/low load area obtained by excluding a part of the high speed area from the low engine load area (including no load). The second area A2 is a low/medium speed/medium load area with a higher load than the first area A1. The fourth area A4 is a low-speed/high-load area in which the load is higher and the rotational speed is lower than in the second area A2. The third area A3 is a middle speed/high load area in which the rotational speed is higher than that of the fourth area A4. The fifth area A5 is a high speed area in which the rotation speed is higher than any of the first to fourth areas A1 to A4.

In the first region A1, partial compression ignition combustion (hereinafter referred to as SPCCI combustion) combining SI combustion and CI combustion is performed. SI combustion is a combustion mode in which the air-fuel mixture is ignited by a spark generated from the spark plug 16, and the air-fuel mixture is forcibly burned by flame propagation that spreads the combustion area from the ignition point to the surroundings. . CI combustion is a combustion mode in which an air-fuel mixture is combusted by self-ignition under an environment of high temperature and high pressure due to compression of the piston 5 . SPCCI combustion, which is a combination of these SI combustion and CI combustion, involves SI combustion of 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. It is a combustion mode in which another air-fuel mixture in the combustion chamber 6 is later subjected to CI combustion by self-ignition (due to further increase in temperature and pressure accompanying SI combustion). “SPCCI” is an abbreviation for “Spark Controlled Compression Ignition”.

SPCCI combustion has the property that heat release during CI combustion is steeper than heat release during SI combustion. As shown in FIG. 8, in the waveform of the heat release rate due to SPCCI combustion, the rising slope at the beginning of combustion corresponding to SI combustion is smaller than the rising slope corresponding to subsequent CI combustion. In other words, the waveform of the heat release rate during SPCCI combustion consists of a first heat release rate portion with a relatively small rising slope based on SI combustion and a second heat release rate portion with a relatively large rising slope based on CI combustion. are formed so as to be continuous in this order. Further, in response to such a tendency of the heat release rate, in SPCCI combustion, the pressure rise rate (dp/dθ) in the combustion chamber 6 occurring during SI combustion is smaller than that during CI combustion.

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 starts. As illustrated in FIG. 8, the slope of the heat release rate waveform changes from small to large at the timing of this self-ignition (that is, the timing at which CI combustion starts). That is, the waveform of the heat release rate in SPCCI combustion has an inflection point (X in FIG. 8) that appears at the timing when CI combustion starts.

After the start of CI combustion, SI combustion and CI combustion are performed in parallel. In CI combustion, the combustion speed of air-fuel mixture is faster than in SI combustion, so the heat generation rate is relatively large. However, since CI combustion is performed after compression top dead center, the slope of the heat release rate waveform does not become excessive. That is, when the compression top dead center is passed, the motoring pressure drops due to the descent of the piston 5, which suppresses the rise in the heat release rate, thereby avoiding an excessive increase in dp/dθ during CI combustion. be. In this way, in SPCCI combustion, CI combustion is performed after SI combustion. ), combustion noise can be suppressed.

SPCCI combustion also ends with the end of CI combustion. Since the CI combustion has a higher combustion speed than the SI combustion, it is possible to advance the combustion end timing as compared to the simple SI combustion (when all the fuel is SI-burned). Therefore, in SPCCI combustion, the end of combustion can be brought close to the compression top dead center in the expansion stroke. As a result, SPCCI combustion can improve fuel efficiency compared to simple SI combustion.

In the first region A1, the above SPCCI combustion is performed in a lean environment (SPCCI_λ>1). That is, the opening degree of the throttle valve 32 is set to an opening degree at which more air than the amount of air corresponding to the stoichiometric air-fuel ratio is introduced into the combustion chamber 6 through the intake passage 30 . Specifically, the processor 60 controls the air-fuel ratio (A/F ) is set to be greater than the stoichiometric air-fuel ratio (14.7), control is executed to cause the air-fuel mixture in the combustion chamber 6 to undergo SPCCI combustion.

Internal EGR that leaves the burned gas in the combustion chamber 6 is performed in many areas of the first area A1. The processor 60 controls the intake VVT 13a and the exhaust VVT 14a to open the intake and exhaust valves 11 and 12 so that a valve overlap is formed in which both the intake and exhaust valves 11 and 12 are opened across the exhaust top dead center. The exhaust valve 12 is opened until it passes the exhaust top dead center (until the beginning of the intake stroke). As a result, the burned gas is drawn back from the exhaust port 10 to the combustion chamber 6, and internal EGR is realized. The period of valve overlap is set so as to achieve an in-cylinder temperature suitable for obtaining a desired SPCCI combustion waveform.

In the second region A2, control is executed to SPCCI burn the air-fuel mixture in an environment where the air-fuel ratio in the combustion chamber 6 substantially matches the stoichiometric air-fuel ratio (SPCCI_λ=1). The opening of the throttle valve 32 is set such that an amount of air corresponding to the stoichiometric air-fuel ratio is introduced into the combustion chamber 6 through the intake passage 30 . In the second operating region A2, the EGR valve 53 is opened to introduce the external EGR gas into the combustion chamber 6. Therefore, in the second operating region A2, the gas air-fuel ratio (G/F), which is the weight ratio of the total gas in the combustion chamber 6 and the fuel, becomes larger than the stoichiometric air-fuel ratio (14.7). Therefore, during operation in the second region A2, control is performed for SPCCI combustion of the air-fuel mixture while forming a G/F lean environment in which the G/F is greater than the stoichiometric air-fuel ratio and the A/F substantially matches the stoichiometric air-fuel ratio. is executed. The opening degree of the EGR valve 53 is set to an opening degree that achieves the stoichiometric air-fuel ratio on an A/F basis.

In the third region A3, control is executed to SPCCI burn the air-fuel mixture in an environment where the A/F in the combustion chamber 6 is slightly richer than the stoichiometric air-fuel ratio (SPCCI_λ≦1). A rich environment is set because a suitable amount of fuel injection is required to handle medium speed and high loads. On the other hand, in the fourth region A4, which is a high-load but low-speed operation region, SPCCI combustion of the air-fuel mixture is performed under an environment where the A/F substantially matches the stoichiometric air-fuel ratio (SPCCI_λ=1). Relatively orthodox SI combustion is performed in the fifth region A5. A/F is set to the stoichiometric air-fuel ratio or a value slightly richer than this (SI_λ≦1). Note that the A/F can be adjusted by the opening of the EGR valve 53 also in these regions A3 to A5.

<During half warm-up>
In the second operation map Q2 during half-warming up, the sixth area B1 corresponds to the combined area of the first and second areas A1 and A2 in the first operation map Q1. The seventh area B2, the eighth area B3 and the ninth area B4 respectively correspond to the third area A3, the fourth area A4 and the fifth area A5 of the first driving map Q1.

In the sixth region B1, as in the second region A2 of the first operation map Q1, control is executed to cause the air-fuel mixture to undergo SPCCI combustion under an environment where the A/F in the combustion chamber 6 substantially matches the stoichiometric air-fuel ratio. (SPCCI_λ=1). In at least a part of the sixth region B1, internal EGR is performed in which the valve overlap period is set and the burned gas remains in the combustion chamber 6. The supercharger 33 is turned on in the relatively high-load region and the relatively high-speed region of the sixth region B1, and is turned off in the other regions. In the sixth region B1, internal EGR is performed, and in a specific region B11, which is a part of the region in which the supercharger 33 is turned off, re-intake correction control, which will be described in detail later, is performed.

The seventh area B2, the eighth area B3, and the ninth area B4 are controlled in the same manner as the third area A3, the fourth area A4, and the fifth area A5 of the first driving map Q1, respectively. That is, in the seventh region B2, the air-fuel mixture is SPCCI-burned in an environment where the A/F in the combustion chamber 6 is slightly richer than the stoichiometric air-fuel ratio (SPCCI_λ≦1). In the eighth region B3, the air-fuel mixture is SPCCI-burned (SPCCI_λ=1) in an environment where the A/F substantially matches the stoichiometric air-fuel ratio. In the ninth region B4, orthodox SI combustion is performed, and A/F is set to the stoichiometric air-fuel ratio or a slightly richer value (SI_λ≦1).

<When cold>
The cold third operation map Q3 consists of only the tenth region C1. In the tenth region C1, control is executed to mix the fuel injected mainly during the intake stroke with air and perform SI combustion. The control in this tenth region C1 is similar to the combustion control of a general gasoline engine.

[About SI rate]
In the present embodiment, SPCCI combustion, which is a combination of SI combustion and CI combustion, is performed in the first to fourth regions A1 to A4 of the first operation map Q1 and the sixth to eighth regions B1 to B3 of the second operation map Q2. executed. In SPCCI combustion, it is important to control the ratio of SI combustion and CI combustion according to operating conditions. In the present embodiment, as the ratio, the SI rate, which is the ratio of the amount of heat release due to SI combustion to the total amount of heat release during one cycle during SPCCI combustion, is used for control. FIG. 8 is also a diagram for explaining this SI rate, and shows changes in the heat release rate (J/deg) depending on the crank angle when SPCCI combustion occurs.

A point X in the waveform of FIG. 8 is an inflection point that appears when the combustion mode switches from SI combustion to CI combustion. The crank angle θci corresponding to this inflection point X is defined as the start timing of CI combustion. The area R1 of the heat release rate waveform located on the advanced side of θci, which is the start timing of CI combustion, is defined as the amount of heat release due to SI combustion, and the area of the heat release rate waveform located on the retarded side of θci. Let R2 be the heat release rate due to CI combustion. This will
(Amount of heat release due to SI combustion)/(Amount of heat release due to SPCCI combustion)
The SI rate defined above can be represented by R1/(R1+R2) using the areas R1 and R2. That is, SI rate=R1/(R1+R2).

In CI combustion, the air-fuel mixture burns simultaneously and multiple times due to self-ignition, so the rate of pressure rise tends to be higher than in SI combustion due to flame propagation. Therefore, if the SI rate is carelessly decreased especially under the condition that the load is high and the fuel injection amount is large, a large amount of noise will be generated. On the other hand, CI combustion does not occur unless the combustion chamber 6 is sufficiently heated and pressurized. Therefore, when the load is low and the fuel injection amount is small, CI combustion does not start until SI combustion has progressed to some extent, and the SI rate inevitably increases.

In consideration of such circumstances, in the present embodiment, a target SI rate, which is a target value of the SI rate, is predetermined for each operating condition of the engine in the operating region where SPCCI combustion is performed. Furthermore, in correspondence with the target SI rate, a target θci, which is the start timing of CI combustion when combustion matching the target SI rate is performed, is also predetermined for each operating condition of the engine.

In order to achieve the target SI rate and target θci described above, control variables such as the timing of main ignition by the spark plug 16, the injection amount/injection timing of fuel from the injector 15, and the EGR rate (external EGR rate and internal EGR rate) are set. It must be adjusted for each operating condition. For example, the more the main ignition timing is advanced, the more fuel is burned by SI combustion, and the SI rate becomes higher. Further, as the fuel injection timing is advanced, more fuel is burned by CI combustion, and the SI rate is lowered. Alternatively, as the in-cylinder temperature increases with an increase in the EGR rate, more fuel is burned by CI combustion, resulting in a lower SI rate. Furthermore, since changes in the SI rate are accompanied by changes in θci, changes in these control variables (main ignition timing, injection timing, EGR rate, etc.) are factors for adjusting θci. In this embodiment, when SPCCI combustion is executed, for example, the main ignition timing, the fuel injection amount/injection timing, the EGR rate (and thus the in-cylinder temperature), etc. are combined to achieve the above-described target SI rate and target θci. controlled as

[Execution mode of internal EGR]
Next, the manner in which internal EGR is executed will be described. In the present embodiment, at least in the sixth region B1 (specific region B11) in the second operation map Q2 during half-warming and in the first region A1 in the first operation map Q1 during warming, the intake valve 11 and the exhaust valve Internal EGR with valve overlap opening both of the 12 valves is performed. The execution mode of internal EGR is the sixth region B1 (SPCCI_λ=1) in which SPCCI combustion is performed by setting the A/F to the stoichiometric air-fuel ratio, and the first region A1 in which SPCCI combustion is performed by setting the A/F to lean. (SPCCI_λ>1).

FIG. 9A is a graph showing an example of the valve overlap period OVL-1 during SPCCI_λ=1 combustion in the sixth region B1. The closing timing EVC of the exhaust valve 12 is set at a crank angle (CA) retarded from TDC, and the opening timing IVO of the intake valve 11 is set at a crank angle advanced from TDC. During the valve overlap period OVL-1, both the intake valve 11 and the exhaust valve 12 are open, and the exhaust pressure is higher than the intake pressure. , return to cylinder 2. In addition, due to the return of the burned gas, the intake air that has once entered the cylinder 2 is blown out from the intake port 9 to the intake passage 30 side. The blown intake air is then sucked into the intake port 9 again. In the operating region of SPCCI_λ=1, the interval between the closing timing EVC of the exhaust valve 12 and the opening timing IVO of the intake valve 11 is set relatively short.

FIG. 9B is a graph showing an example of the valve overlap period OVL-2 during SPCCI_λ>1 combustion in the first region A1. Compared to the SPCCI_λ=1 combustion in FIG. 9A, the closing timing EVC of the exhaust valve 12 is retarded and the opening timing IVO of the intake valve 11 is advanced. As a result, the valve overlap period OVL-2 is considerably longer than the valve overlap period OVL-1. This is because when SPCCI combustion is performed in a lean environment, ignitability deteriorates, so it is necessary to increase the internal EGR amount to raise the cylinder temperature.

FIG. 9C is a graph showing the relationship between the return amount of burned gas (internal EGR amount) and the valve overlap time. The return amount tends to increase as the valve overlap time increases. FIG. 9C shows the return amount when SPCCI_λ=1 combustion in FIG. 9A and the return amount when SPCCI_λ>1 combustion in FIG. pointed out. Due to the difference between the valve overlap periods OVL-1 and OVL-2, the return amount during SPCCI_λ>1 combustion is considerably larger. OVL-2 may be set to approximately +30 degrees in crank angle CA compared to OVL-1. In this case, the burned gas that has returned into the cylinder 2 from the exhaust port 10 may even be blown out from the intake port 9 .

The degree of variation in the air-fuel ratio among the #1 to #4 cylinders 2 may differ depending on the return amount of the burned gas. As described above, in this embodiment, the surge tank 36 of the intake passage 30 employs a center intake structure, and the independent intake passage 37 is structurally characterized by being extremely short. On the other hand, the exhaust manifold 42 of the exhaust passage 40 has an end unevenly distributed exhaust structure in which the #1 cylinder 2 has the shortest exhaust passage length.

At the time of SPCCI_λ=1 combustion in which the closing timing EVC of the exhaust valve 12 is relatively early, the exhaust valve 12 is closed. For this reason, the degree of re-intake of the intake air, which is exclusively derived from the center intake structure on the intake side, becomes a factor of variations in the air-fuel ratio among the cylinders 2 . On the other hand, during SPCCI_λ>1 combustion in which the valve closing timing EVC is relatively late, the return amount of the burned gas increases significantly. For this reason, the degree of return of the burned gas due to the shape feature of the unevenly-distributed exhaust structure on the exhaust side becomes a factor of air-fuel ratio variation among the cylinders 2 .

[Behavior of intake and exhaust during execution of internal EGR]
FIG. 10 is a schematic diagram for explaining the state of the return of exhaust gas to the cylinder, the return of intake air, and re-intake due to the execution of internal EGR. The intake air G1 is introduced into the surge tank 36 from the third passage 303 of the intake passage 30 . The exhaust gas G2 after combustion is sent out to the exhaust pipe 43 through the exhaust manifold 42 . On the other hand, when internal EGR is executed, return exhaust gas G3 is generated in which part of the exhaust gas G2 returns to each of the #1 to #4 cylinders 2 due to the setting of the valve overlap period and the intake/exhaust differential pressure. Then, return intake air G4 is generated in which the intake air G1 once introduced into each of the #1 to #4 cylinders 2 blows out in the form of being pushed out by the return exhaust gas G3.

First, the behavior of the return intake G4 will be described. Since the independent intake passages 37_#1 to #4 corresponding to the #1 to #4 cylinders 2 are short, the return intake air G4 can flow over the independent intake passages 37_#1 to #4 to the surge tank 36. be. In FIG. 10, symbols g11, g12, g13, and g14 schematically indicate blown intake air returning to the surge tank 36 from the independent intake passages 37_#1, 37_#2, 37_#3, and 37_#4, respectively.

If the independent intake passages 37_#1-#4 are appropriately long, the return intake air G4 is blown into each of the independent intake passages 37_#1-#4 and then re-inhaled into each intake port 9. FIG. Therefore, there is no variation in the air-fuel ratio among the #1 to #4 cylinders. However, if the return intake air G4 blows out to the internal space 36R of the surge tank 36, the air-fuel ratio will vary. Specifically, in the center air intake structure as in this embodiment, among the four cylinders 2 arranged in a line, the #1 and #4 cylinders 2 (first cylinder/end cylinder) located at both ends and the # A variation in the air-fuel ratio occurs between the #2 and #3 cylinders 2 (second cylinder/central side cylinder) sandwiched between the #1 and #4 cylinders 2.

Focus on the intake air g11 blown from #1 cylinder 2. FIG. As the internal combustion cycle progresses thereafter, most of the blown intake air g11 returns to the independent intake passage 37_#1 from which it was blown, and is reinhaled into the #1 cylinder 2 through the intake port 9_#1, as indicated by the arrow r11. . However, since it is a center intake structure, it is also affected by the flow tendency of intake air G1 that tends to flow toward #2 and #3 cylinders 2 on the center side. It floats in the inner space 36R of 36. Then, as indicated by arrows r11a and r11b, part of the blown-out intake air g11 enters the adjacent independent intake passage 37_#2 and heads toward the other independent intake passages 37_#3 and #4. Due to such behavior, the blown intake air g11 does not completely return to #1 cylinder 2, and as a result, the amount of air in #1 cylinder 2 becomes insufficient.

The same applies to the blown intake air g14 from #4 cylinder 2. Most of the blown intake air g14 blown into the internal space 36R returns to the independent intake passage 37_#4 and is re-inhaled into #4 cylinder 2 through the intake port 9_#4. However, part of the blown intake air g14 enters the adjacent independent intake passages 37_#3 or goes toward the other independent intake passages 37_#1 and #2, as indicated by arrows r14a and r14b. Due to this behavior, #4 cylinder 2 also runs short of air.

On the other hand, in #2 and #3 cylinders 2, the amount of air becomes excessive. That is, the intake air g12 blown out from the #2 cylinder 2 is also assisted by the flow of the intake air G1, and returns to the independent intake passage 37_#2 from which almost the entire amount was blown out, as indicated by the arrow r12, to the intake port 9_#2. It is reinhaled into #2 cylinder 2 through. Similarly, the intake air g13 blown from #3 cylinder 2 also returns to the independent intake passage 37_#3 from which almost the entire amount was blown, and is reinhaled into #3 cylinder 2 through the intake port 9_#3, as indicated by arrow r13. be.

In addition, part of the intake air g11 blown from the #1 cylinder 2 (arrow r11a) and part of the intake air g14 blown from the #4 cylinder 2 (arrow r14b) are drawn into the independent intake passage 37_#2. obtain. Similarly, part of the intake air g14 blown from the #4 cylinder 2 (arrow r14a) and part of the intake air g11 blown from the #1 cylinder 2 (arrow r11b) can be taken into the independent intake passage 37_#3. . Therefore, the re-intake amount of the intake ports 9_#2 and #3 of the #2 and #3 cylinders 2 becomes larger than the original return intake G4 amount. Due to the above behavior, the re-intake amount (predetermined amount) of intake air in #1 and #4 cylinders 2 becomes larger than the amount of re-intake in cylinders #2 and #3.

The behavior of the return intake air G4 as described above does not manifest itself in an operating state where the intake pressure is relatively high (the intake pressure difference is small). For example, when the supercharger 33 is operating to increase the intake pressure, the return intake air G4 does not reach the surge tank 36 even if the independent intake passage 37 is short. When the EGR valve 53 is opened and the external EGR enters the surge tank 36, the external EGR enters each cylinder 2 so as to cancel out the variation in the re-intake amount. Variation in the air-fuel ratio between two #4 cylinders does not occur.

Next, the behavior of the return exhaust gas G3 will be described. In the exhaust manifold 42, as shown in FIG. 5, the #1 cylinder 2 (fourth cylinder) independent exhaust pipe 44_#1 has the shortest route to the exhaust pipe 43, and the #2 to #4 cylinders 2 (third cylinder) have the shortest path. The independent exhaust pipes 44_#2 to 44_#4 of the cylinders) have an uneven end exhaust structure in which the path to the exhaust pipe 43 gradually becomes longer. When internal EGR is performed, the blown exhaust gas g2, which is the burned gas that has been blown out to the confluence portion 45 of the exhaust manifold 42 or the exhaust pipe 43, is returned as the return exhaust gas G3, and the independent exhaust pipes 44_#1 to 44_#4 and each Through exhaust ports 10_#1 to #4, they return to #1 to #4 cylinders 2, respectively.

The independent exhaust pipe 44_#1 communicates linearly with the exhaust pipe 43 and has the shortest path length. It becomes 4 cylinder 2. The independent exhaust pipes 44_#1-#4 are not as short as the independent intake passages 37_#1-#4 on the intake side. For this reason, in the case of SPCCI_λ=1 combustion in which the internal EGR amount is small and the amount of burned gas returned is small, this does not occur. As a result, variations in G/F occur between #1 to #4 cylinders 2. Naturally, the re-intake amount of burned gas in #1 cylinder 2 (one end cylinder) tends to be large, and the re-intake amount tends to be small in #2 to #4 cylinders 2 (remaining three cylinders).

[Inherent rebreathing correction amount]
FIGS. 11A to 11D are graphs showing correction tendencies of the fuel injection amounts of #1 to #4 cylinders 2 in the operating region of SPCCI_λ=1. The BGR ratio on the horizontal axis of these graphs is an index indicating the remaining amount of burned gas in the cylinder 2 (combustion chamber 6), and the greater the BGR ratio, the greater the amount of burned gas (that is, the amount of internal EGR). indicate. "+" on the vertical axis indicates correction for increasing the target fuel injection amount, and "-" indicates correction for decreasing the target fuel injection amount.

The correction amount of the target fuel injection amount shown in the above graph is a specific correction amount determined by the center intake structure employed in this embodiment. That is, the correction amount is a specific re-intake correction amount set for each of #1 to #4 cylinders 2 according to the re-intake amount of intake air by internal EGR in each cylinder 2 shown in FIG. The first correction unit 65 (FIG. 6) primarily corrects the target fuel injection amount set for #1 to #4 cylinders 2 based on these re-intake correction amounts. The storage unit 68 stores table data of the reinhalation correction amount corresponding to the graph.

As described above, the re-intake amount of intake air is relatively small for #1 and #4 cylinders 2 (first cylinder/end cylinder). On the other hand, for #2 and #3 cylinders 2 (second cylinders/central side cylinders), the reintake amount of intake air is relatively large. In this case, if the #1 to #4 cylinders 2 are caused to inject fuel according to the target fuel injection amount, the air amount in the #1 and #4 cylinders 2 becomes insufficient and the A/F becomes rich. In the #3 cylinder 2, the amount of air becomes excessive and the A/F becomes lean, resulting in variations in the A/F. In order to correct this variation, the target fuel injection amount is adjusted so that the A/F for #1 and #4 cylinders 2 is on the lean side and the A/F for #2 and #3 cylinders 2 is on the rich side. Correction should be made.

Specifically, as shown in FIGS. 11A and 11D, for #1 and #4 cylinders 2, a re-intake correction amount for reducing the target fuel injection amount is set. As a result, the #1 and #4 cylinders 2, in which the A/F is rich due to insufficient re-intake amount, can be brought closer to λ=1. Also, the degree of decrease in the target fuel injection amount is set to increase as the BGR ratio increases. This is because the larger the internal EGR amount, the larger the amount of return intake air G4 that is blown out to the surge tank 36, so that the variation in A/F between the cylinders 2 tends to increase.

On the other hand, as shown in FIGS. 11B and 11C, for #2 and #3 cylinders 2, a re-intake correction amount for increasing the target fuel injection amount is set, and the degree of increase is determined by the BGR ratio. It is set to increase as it increases. By performing such a correction, it is possible to bring λ=1 closer to #2 and #3 cylinders 2 where the re-intake amount is excessive and the A/F is lean.

FIGS. 12A to 12D are graphs showing correction tendencies of the fuel injection amounts of #1 to #4 cylinders 2 in the operating region of SPCCI_λ>1. The correction amount of the target fuel injection amount shown in this graph is a specific correction amount determined by the end unevenly distributed exhaust structure of the exhaust manifold 42 employed in this embodiment. The first correction unit 65 primarily corrects the target fuel injection amount set for the #1 to #4 cylinders 2 based on these unique correction amounts. The storage unit 68 stores correction amount table data corresponding to the graph.

As described above, for #1 cylinder 2 (fourth cylinder/one end cylinder), the amount of re-intake of burned gas from the exhaust port 10 is relatively large. On the other hand, for #2 to #4 cylinders 2 (third cylinder/remaining three cylinders), the re-intake amount of burned gas is relatively small. In this case, if the #1 to #4 cylinders 2 are caused to inject fuel according to the target fuel injection amount, the amount of gas including air becomes excessive in the #1 cylinder 2, causing the G/F to become lean. ∼ #4 Cylinder 2 lacks the gas amount and the G/F becomes rich, resulting in variations in G/F. To correct this variation, the target fuel injection amount should be corrected so that the G/F of #1 cylinder 2 is on the rich side and the G/F of #2 to #4 cylinder 2 is on the lean side. good.

Specifically, as shown in FIG. 12A, for #1 cylinder 2, a correction amount for increasing the target fuel injection amount is set. As a result, the G/F of #1 cylinder 2, whose G/F fluctuates to the rich side due to insufficient re-intake, can be brought closer to the target G/F. Further, the degree of increase of the target fuel injection amount is set so as to increase as the BGR ratio increases. This is because the amount of return exhaust gas G3 from the exhaust pipe 43 increases as the amount of internal EGR increases, and the variation in G/F between the cylinders 2 tends to increase.

On the other hand, as shown in FIGS. 12B, 12C, and 12D, for #2, #3, and #4 cylinders 2, a correction amount for decreasing the target fuel injection amount is set. The degree is set to increase as the BGR ratio increases. By performing such correction, it is possible to bring the G/F of #2, #3, and #4 cylinders 2 in which the amount of return exhaust gas G3 is excessive and the G/F is lean closer to the target G/F.

[Correction of unique rebreathing correction amount]
In the present embodiment, the second correction unit 66 (FIG. 6) adjusts the reintake correction amount unique to each cylinder 2 shown in at least FIGS. to correct further. The operating conditions and environmental conditions are, for example, intake pressure, valve overlap amount, closing timing of the exhaust valve 12, engine speed, atmospheric pressure, and exhaust gas temperature. These are the variables that affect the amount of burnt gas returned to cylinder 2 by internal EGR. When the return amount of burned gas fluctuates, the amount of intake air blown out from the intake port 9 also fluctuates. For example, the amount of blown intake air g11 to g14 (see FIG. 10) entering the surge tank 36 fluctuates. Therefore, the mode of reinhalation of blown intake air g11 to g14 may also vary. Therefore, it is desirable to perform secondary correction of the reinhalation correction amount according to the above conditions, instead of setting the reinhalation correction amount in a fixed manner. As a result, the A/F between two cylinders #1 to #4 can be aligned with higher accuracy.

Hereinafter, with reference to FIGS. 13 and 14, a specific description will be given of aspects of the secondary correction performed by the second correction unit 66 for each of the above conditions. FIGS. 13A to 13C and FIGS. 14A to 14C are graphs showing correction tendencies of specific rebreathing correction amounts for variables dependent on operating conditions and environmental conditions. Each graph shows "Ref" which means a reference value. When each variable is the reference value Ref, the secondary correction amount is multiplied by 1, and as a result, the unique reinhalation correction amount value set by the first correction unit 65 is maintained. The region where the reference value Ref is higher than the secondary correction amount is a region where the specific reinhalation correction amount is increased, and the region where the reference value Ref is lower is a region where the specific reinhalation correction amount is decreased. For example, assume that the re-intake correction amount specific to a certain cylinder 2 is a correction value that increases the target fuel injection amount by +3% at a certain internal EGR amount. In this case, when the variable is the reference value Ref, the +3% increase is maintained, and in a region higher than the secondary correction amount of the reference value Ref, +3% is increased to +3.5%, and the secondary correction amount of the reference value Ref is increased. For example, +3% is reduced to +2.5% in a region lower than the correction amount.

<Intake pressure>
FIG. 13A is a graph showing the relationship between the intake pressure and the secondary correction amount of the specific re-intake correction amount. As shown in this graph, the second correction unit 66 corrects the re-intake correction amount to decrease as the intake pressure increases. That is, when the intake pressure is higher than the reference value Ref, the re-intake correction amount is decreased, and when the intake pressure is lower than the reference value Ref, the re-intake correction amount is increased. The intake pressure can be grasped from the measured values of the first intake pressure sensor SN6 and the second intake pressure sensor SN8.

The amount of burned gas returned to the cylinder 2 by internal EGR depends on the magnitude of the differential pressure between intake and exhaust, which is the difference between the intake pressure and the exhaust pressure. When the intake/exhaust differential pressure is large, the return amount of high pressure side exhaust gas (burnt gas) increases, and intake air is blown out of the cylinder 2 by that amount. When the intake pressure becomes relatively high with respect to the exhaust pressure, the intake/exhaust differential pressure becomes smaller, so the amount of intake air blown out from the intake port 9 decreases, and the amount of re-intake of the intake air naturally decreases. Therefore, the variation in the re-intake amount among the #1 to #4 cylinders 2 is also reduced. Therefore, by correcting such that the re-intake correction amount is reduced as the intake pressure increases, correction of the target fuel injection amount that is in line with the actual situation can be achieved. In order to simplify the control, if the intake pressure exceeds a predetermined threshold value or if the intake/exhaust differential pressure falls below a predetermined threshold value, the re-intake correction itself is not executed. can be

As shown in FIGS. 11A and 11D, for #1 and #4 cylinders 2, a re-intake correction amount for decreasing the target fuel injection amount is set as primary correction. For these #1 and #4 cylinders 2, secondary correction is performed to reduce the re-intake correction amount so that the reduction degree of the target fuel injection amount is reduced in the region where the intake pressure is higher than the reference value Ref. For example, assume that the re-intake correction amount specific to #1 cylinder 2 is a correction value that reduces the target fuel injection amount by -3% at a certain internal EGR amount. In this case, in a region where the intake pressure is higher than the reference value Ref, the -3% decrease is secondarily corrected to, for example, a -2.5% decrease. Conversely, in a region where the intake pressure is lower than the reference value Ref, the -3% decrease is secondarily corrected to, for example, a -3.5% decrease. When the intake pressure is the reference value Ref, the -3% reduction is maintained.

On the other hand, as shown in FIGS. 11B and 11C, for #2 and #3 cylinders 2, a re-intake correction amount for increasing the target fuel injection amount is set as primary correction. For #2 and #3 cylinders 2, secondary correction is performed to increase the re-intake correction amount so as to decrease the degree of increase in the target fuel injection amount in a region where the intake pressure is higher than the reference value Ref. For example, assume that the re-intake correction amount unique to #2 cylinder 2 is a correction value that increases the target fuel injection amount by +3% at a certain internal EGR amount. In this case, in a region where the intake pressure is higher than the reference value Ref, the +3% increase is secondarily corrected to, for example, a -2.5% increase. Conversely, in a region where the intake pressure is lower than the reference value Ref, the +3% increase is secondarily corrected to, for example, a +3.5% increase. When the intake pressure is the reference value Ref, the +3% reduction is maintained. Such secondary correction for each cylinder 2 is the same for the following variables (description is omitted below).

<Valve overlap amount>
FIG. 13B is a graph showing the relationship between the valve overlap amount and the secondary correction amount of the specific re-intake correction amount. As shown in this graph, the second correction unit 66 corrects so that the re-intake correction amount is increased as the valve overlap amount increases. That is, when the valve overlap amount is larger than the reference value Ref, the reintake correction amount is increased, and when the valve overlap amount is smaller than the reference value Ref, the reintake correction amount is decreased. The valve overlap amount can be grasped from the measured values of the intake cam angle sensor SN12 and the exhaust cam angle sensor SN13.

Assuming that the engine speed is constant, the return amount of the burned gas to the cylinder 2 increases as the valve overlap amount increases. In other words, the period in which both the intake valve 11 and the exhaust valve 12 are open becomes longer, so the burned gas is more likely to return to the cylinder 2 . Along with this, the amount of intake air blown out from the intake port 9 also increases, so that the variation in the re-intake amount of the intake air among the #1 to #4 cylinders 2 also tends to increase. Therefore, by increasing the re-intake correction amount as the valve overlap amount increases, it is possible to suppress variations in A/F among the cylinders #1 to #4.

<Exhaust valve closing timing>
FIG. 13C is a graph showing the relationship between the closing timing of the exhaust valve 12 and the secondary correction amount of the inherent re-intake correction amount. As shown in this graph, the second correction unit 66 corrects so that the re-intake correction amount is increased as the closing timing of the exhaust valve 12 is delayed. That is, when the closing timing of the exhaust valve 12 is retarded from the reference closing timing (Ref), the re-suction correction amount is increased to advance the closing timing of the exhaust valve 12 from the reference closing timing (Ref). When the reinhalation correction amount is decreased, a correction is performed. The closing timing of the exhaust valve 12 can be grasped, for example, from the measured value of the exhaust cam angle sensor SN13.

As the closing timing of the exhaust valve 12 is delayed, the amount of burned gas returned to the cylinder 2 is increased. That is, by retarding the closing timing of the exhaust valve 12, the period in which both the intake valve 11 and the exhaust valve 12 are open becomes longer, so the burned gas that has once exited the exhaust port 10 returns to the cylinder 2. becomes easier. Along with this, the amount of intake air blown out from the intake port 9 also increases, so that the variation in the re-intake amount of the intake air among the #1 to #4 cylinders 2 tends to increase. Therefore, by increasing the re-intake correction amount as the closing timing of the exhaust valve 12 is retarded, the variation in A/F among the #1 to #4 cylinders 2 can be suppressed.

<Engine speed>
FIG. 14A is a graph showing the relationship between the engine speed and the secondary correction amount of the specific re-suction correction amount. As shown in this graph, the second correction unit 66 corrects the re-suction correction amount to decrease as the engine speed increases. That is, when the engine speed is higher than the reference value Ref, the re-suction correction amount is decreased, and when the engine speed is lower than the reference value Ref, the re-suction correction amount is increased. will be The engine speed can be obtained, for example, from the measured value of the crank angle sensor SN1.

Assuming that the valve overlap amount is constant, as the engine speed increases, the valve overlap time, which is the time during which both the intake valve 11 and the exhaust valve 12 are actually open, becomes shorter, and the cylinder of the burnt gas The amount of return to 2 is reduced. Along with this, the amount of intake air blown out from the intake port 9 also decreases, so the variation in the amount of reintake of the intake air among the #1 to #4 cylinders 2 also tends to decrease. Therefore, by correcting so as to reduce the re-intake correction amount as the engine speed increases, it is possible to appropriately suppress variations in A/F among the #1 to #4 cylinders.

<Atmospheric pressure>
FIG. 14B is a graph showing the relationship between the atmospheric pressure and the secondary correction amount of the inherent re-inhalation correction amount. As shown in this graph, the second correction unit 66 corrects so that the reinhalation correction amount increases as the atmospheric pressure increases. That is, when the atmospheric pressure is higher than the reference value Ref, the re-suction correction amount is increased, and when the atmospheric pressure is lower than the reference value Ref, the re-suction correction amount is decreased. The atmospheric pressure can be grasped from the measured value of the atmospheric pressure sensor SN15.

As the atmospheric pressure rises, the exhaust pressure rises. On the other hand, since the intake pressure depends on the opening of the throttle valve 32, the boost pressure, etc., it is less likely to be affected by the atmospheric pressure. For this reason, the differential pressure between intake and exhaust air increases as the atmospheric pressure rises. Conversely, if the engine body 1 is placed in an environment where the atmospheric pressure is low, such as when traveling at high altitudes, the differential pressure between intake and exhaust will decrease. As the intake/exhaust differential pressure increases, the amount of burned gas returned to the cylinder 2 also increases. Along with this, the amount of intake air blown out from the intake port 9 also increases, so that the variation in the re-intake amount of the intake air among the #1 to #4 cylinders 2 tends to increase. Therefore, by increasing the re-intake correction amount as the atmospheric pressure increases, it is possible to appropriately suppress variations in A/F among the #1 to #4 cylinders.

<Exhaust gas temperature>
FIG. 14B is a graph showing the relationship between the exhaust gas temperature and the secondary correction amount of the specific re-intake correction amount. As shown in this graph, the second correction unit 66 corrects the reinhalation correction amount so that it increases as the exhaust gas temperature increases. That is, when the exhaust gas temperature is higher than the reference value Ref, the re-suction correction amount is increased, and when the exhaust gas temperature is lower than the reference value Ref, the re-suction correction amount is decreased. The exhaust gas temperature can be grasped from the measured value of the exhaust temperature sensor SN10.

As the exhaust gas temperature rises, the exhaust pressure rises. On the other hand, the intake pressure is less affected by the atmospheric pressure as described above. Therefore, the differential pressure between intake and exhaust is increased. As the intake-exhaust differential pressure increases, the amount of burned gas returned to the cylinder also increases. Along with this, the amount of intake air blown out from the intake port also increases, so the variation in the re-intake amount of the intake air among the #1 to #4 cylinders 2 also tends to increase. Therefore, by increasing the re-intake correction amount as the exhaust gas temperature rises, it is possible to appropriately suppress variations in A/F among the #1 to #4 cylinders.

The above is an example of the secondary correction of the specific re-suction correction amount (FIG. 11) in the operating region of SPCCI_λ=1. Similarly, secondary correction may also be performed for the specific reinhalation correction amount (FIG. 12) in the operating region of SPCCI_λ>1. However, in the combustion of SPCCI_λ>1, the ratio of the air amount to the fuel is considerably larger than that of λ=1. /F (G/F) variation does not appear so much. Therefore, in the operating region of SPCCI_λ>1, the secondary correction of the specific re-suction correction amount can be omitted.

[Fuel injection amount determination process]
As shown in FIGS. 11A to 11D, the specific reinhalation correction amount is predetermined according to the BGR ratio. Since the BGR ratio corresponds to the internal EGR amount, it is necessary to grasp the internal EGR amount in the current operating state in order to set the re-intake correction amount. Since there is no sensor that directly measures the internal EGR amount, it is necessary to obtain the internal EGR amount by calculation with reference to the sensor values of some of the sensors SN1 to SN15.

In this embodiment, the EGR prediction unit 64 calculates a predicted value of the internal EGR amount according to the operating state. FIG. 15 is a schematic diagram showing the process of obtaining the predicted value of the internal EGR amount. EGR prediction section 64 includes a polynomial model calculation section 641 and a multiplier 642 . The polynomial model calculation unit 641 applies the polynomial model stored in the storage unit 68 to calculate the provisional value of the internal EGR amount. The polynomial model is a polynomial model whose elements are the amount of valve overlap, the closing timing of the exhaust valve, the differential pressure between intake and exhaust, and the engine speed. A multiplier 642 multiplies the provisional value of the internal EGR amount by a coefficient determined by the exhaust gas temperature and a coefficient determined by the atmospheric pressure to calculate a predicted value of the internal EGR amount.

The polynomial model calculation unit 641 receives as data the opening/closing timings of the intake valve 11 and the exhaust valve 12, the intake pressure, the exhaust pressure, the combined value of the BGR coefficient and the engine speed. Data on the opening/closing timings of the intake valve 11 and the exhaust valve 12 are given based on the measured values of the intake cam angle sensor SN12 and the exhaust cam angle sensor SN13. The intake pressure is given based on the measured values of the first intake pressure sensor SN6 and the second intake pressure sensor SN8. The exhaust pressure is given as a calculated value obtained from, for example, the exhaust amount estimated from the intake air amount measured by the air flow sensor SN4 and the exhaust gas temperature measured by the exhaust temperature sensor SN10. The BGR coefficient is a coefficient that is predetermined in association with the BGR ratio and the engine speed. The engine speed is given based on the measured value of the crank angle sensor SN1.

The polynomial model calculator 641 derives the valve overlap amount from the opening timing of the intake valve 11 and the closing timing of the exhaust valve 12 among the opening/closing timing data of the intake valve 11 and the exhaust valve 12 . Also, the closing timing of the exhaust valve 12 is acquired from the opening/closing timing data of the exhaust valve 12 . A polynomial model calculator 641 calculates an intake-exhaust differential pressure from the intake pressure and the exhaust pressure. Furthermore, the polynomial model calculator 641 identifies the BGR coefficient determined by the engine speed from the input BGR coefficient and the current engine speed. This BGR coefficient is a coefficient corresponding to the valve overlap time.

The valve overlap amount, the closing timing of the exhaust valve 12, the intake/exhaust differential pressure, and the engine speed (valve overlap time) are all the return amount of the burned gas discharged from the exhaust port 10 to the cylinder 2, that is, It is a variable closely related to the internal EGR amount. The polynomial model calculator 641 uses a polynomial regression model represented by polynomials of these variables to provisionally derive an estimated value of the internal EGR amount.

The reason why the value output by the polynomial model calculation unit 641 is the provisional value is that the predicted value of the internal EGR amount is derived by further considering the atmospheric pressure and the exhaust gas temperature, which may affect the return amount of the burned gas. It's for. A multiplier 642 multiplies the internal EGR amount provisional value output from the polynomial model calculation unit 641 by a predetermined coefficient associated with the exhaust gas temperature and a predetermined coefficient associated with the atmospheric pressure. A predicted value of the EGR amount is calculated.

FIG. 16 is an explanatory diagram showing a procedure for correcting the target fuel injection amount (step S0) set by the injection setting unit 63 according to the operating state such as the degree of opening of the accelerator. First, the EGR prediction unit 64 calculates the predicted value of the internal EGR amount by the above method (step S1). Next, as shown in FIGS. 11A to 11D, the first correction unit 65 sets a unique reintake correction amount determined by the engine structure for each of #1 to #4 cylinders 2 ( step S2).

Subsequently, the second correction unit 66 determines a correction coefficient for secondary correction of the specific reinhalation correction amount set by the first correction unit 65 with reference to variables based on the operating state (step S3). The variables are, as shown in FIGS. 13(A) to (C) and FIGS. 14(A) to (C), the intake pressure, the amount of valve overlap, the closing timing of the exhaust valve 12, the engine speed, and the atmospheric pressure. and exhaust gas temperature. Then, the second correction unit 66 corrects the target fuel injection amount (step S0) set based on the accelerator opening for each of #1 to #4 cylinders 2 using the determined correction coefficient (step S4). .

[Combustion control flow]
FIG. 17 is a flow chart showing a specific example of combustion control by the processor 60 of this embodiment. The fuel injection control unit 61 reads various signals from sensors SN1 to SN15 shown in FIG. 6 and other sensors, and acquires information about the operating state of the engine body 1 and environmental information that affects combustion conditions (step S11). .

Subsequently, the injection setting unit 63 determines the target fuel injection amount according to the accelerator depression amount detected by the accelerator opening sensor SN14 (step S12). Then, based on the information obtained in step S1, the driving state determination unit 62 identifies which region of the driving maps Q1 to Q3 shown in FIGS. 7A to 7C corresponds to the current driving point. (Step S13). Although there are a wide variety of operating points, the options for re-inhalation correction are simply shown here. That is, the operating point is, for example, the execution region of SPCCI combustion (SPCCI_λ>1) under a lean environment adopted in the first region A1 of the first operation map Q1, or, for example, the It is determined whether the SPCCI combustion (SPCCI_λ=1) is executed under an environment substantially matching the theoretical air-fuel ratio adopted in the sixth region B1 (specific region B11).

If the operating region is SPCCI_λ=1 (NO in step S13), the EGR prediction unit 64 uses a polynomial model to calculate the predicted value of the internal EGR amount for each of #1 to #4 cylinders 2 ( step S14). Next, the re-intake correction amount for each cylinder 2 is determined according to the internal EGR amount for each of #1 to #4 cylinders 2 (step S15). In the operating region of SPCCI_λ=1, the resuction correction amount is determined in two steps. Specifically, as described with reference to FIG. 16, the first correction unit 65 sets a unique re-intake correction amount for each of #1 to #4 cylinders 2 according to the predicted value of the internal EGR amount (see FIG. 16 step S2). Next, the second correction unit 66 refers to the variable based on the operating state, performs secondary correction on the reinhalation correction amount set by the first correction unit 65, and sets a new reinhalation correction amount (Fig. 16 step S3). Using this new re-intake correction amount, the injection setting unit 63 corrects the target fuel injection amount set in step S12 (step S16).

Subsequently, in order to realize good SPCCI combustion, the fuel injection control unit 61 determines the target θci (step S17). As described above, the target θci is the crank angle corresponding to the inflection point X that appears when the combustion mode switches from SI combustion to CI combustion (see FIG. 8). After that, the ignition control unit 67 determines the ignition timing of the spark plug 16 for realizing the determined target θci (step S18). Then, the ignition control unit 67 causes the ignition plug 16 to ignite at the determined ignition timing (step S19).

The operation region of SPCCI_λ>1 (YES in step S13) is the same as the case of SPCCI_λ=1 except for the re-suction correction amount setting step (step S22). The EGR prediction unit 64 uses a polynomial model to find the predicted value of the internal EGR amount for each of #1 to #4 cylinders 2 (step S21). Next, the re-intake correction amount for each cylinder 2 is determined according to the internal EGR amount for each cylinder #1 to #4 (step S22). In the operating region of SPCCI_λ>1, the secondary correction is not performed, and the unique reinhalation correction amount shown in FIG. 12 is adopted as the correction coefficient as it is. That is, the first correction unit 65 sets a unique re-intake correction amount for each of #1 to #4 cylinders 2 according to the predicted value of the internal EGR amount.

Using the re-intake correction amount set in step S22, the injection setting unit 63 corrects the target fuel injection amount set in step S12 (step S23). Subsequently, the fuel injection control unit 61 determines the target θci for SPCCI combustion (step S24). After that, the ignition control unit 67 determines the ignition timing of the spark plug 16 for realizing the determined target θci (step S25), and causes the spark plug 16 to perform the ignition operation at that ignition timing (step S26).

[Effect]
According to the engine combustion control system according to the present embodiment described above, even if the re-intake amount of intake air due to internal EGR varies among the plurality of cylinders 2 for some reason, The target fuel injection amount is corrected. For example, even if there is a unique difference in the reintake amount between the cylinders 2 due to the structural characteristics of the intake passage, the air-fuel ratio can be adjusted for each cylinder 2 by the reintake correction amount set by the first correction unit 65. can. As a result, it becomes possible to match the air-fuel ratios among the plurality of cylinders 2 .

Further, the second correction unit 66 secondarily corrects the re-intake correction amount set by the first correction unit 65 according to the engine speed. Engine speed is proportional to the time that valve overlap is currently occurring. In other words, whether the engine speed is high or low leads to the length of the valve overlap time. Therefore, the engine speed affects the amount of burned gas returned to the cylinder by internal EGR, and further affects the amount of intake air blown out from the intake port. If the blown intake air amount fluctuates, the reinhalation mode may also fluctuate. Therefore, the air-fuel ratio for each cylinder 2 can be adjusted more accurately by correcting the re-intake correction amount according to the engine speed instead of setting it fixedly. As a result, it is possible to accurately match the air-fuel ratios among the plurality of cylinders by reflecting the operating conditions.

Further, the second correction unit 66 adjusts the re-intake correction amount according to variables such as the intake pressure, valve overlap amount, closing timing of the exhaust valve 12, atmospheric pressure, and exhaust gas temperature, in addition to the engine speed. Secondary correction. These variables also affect the amount of burned gas returned to the cylinder by internal EGR, and also affect the amount of intake air blown out from the intake port 9 . Therefore, the air-fuel ratio of each cylinder 2 can be adjusted more accurately by the second correction unit 66 performing secondary correction of the re-intake correction amount in consideration of these variables.

In particular, the independent intake passage 37 of this embodiment is set short enough to allow the intake air blown out from the intake port 9 to reach the surge tank 36 during the period of valve overlap. For this reason, while the vehicle-mountability of the engine body 1 can be improved, for example, the intake air once discharged from the intake port 9 of the #1 cylinder 2 is not reinhaled into itself, and passes through the surge tank 36 to the #2 cylinder 2. The phenomenon of re-inhalation is likely to occur. However, even if such re-intake imbalance occurs between the cylinders 2, the target fuel injection amount is corrected for each cylinder 2 by the re-intake correction amount executed by the first correction unit 65 and the second correction unit 66. Therefore, variations in the air-fuel ratio between the cylinders 2 can be suppressed.

[Modified embodiment]
Although the embodiments of the present invention have been described above, the present invention is not limited to the above-described embodiments, and for example, the following modified embodiments are possible.

(1) In the above embodiment, it is assumed that the independent intake passage 37 is extremely short, that is, the intake passage 30 has an extremely short length, and the reintake amount of the intake air is inherently varied. This is just an example, and for example, due to the difference in path length of the independent intake passage 37 between #1 to #4 cylinders 2, or due to other factors (arrangement of valves, etc.), unique It can also be applied to the case where the re-inhaled amount varies.

(2) In the above-described embodiment, the third passage 303, which is the intake passage immediately upstream of the surge tank 36, introduces the intake air to the central region in the longitudinal direction of the surge tank 36 as an example. The mode of intake air introduction to the surge tank 36 is not limited to the center intake structure. For example, as in the exhaust passage 40 (FIG. 5), an end uneven intake structure may be employed in which intake air is introduced from a position biased toward the #1 cylinder 2 side or the #4 cylinder 2 side. In this case, the target fuel injection amount may be corrected so that the air-fuel ratio becomes richer on the side closer to #1 cylinder 2 or #4 cylinder 2, and the air-fuel ratio on the far side toward the leaner side.

(3) In the above embodiment, the channel length of the exhaust manifold 42 is the shortest in the #1 cylinder 2, and the channel length becomes longer toward the ends of the cylinders #2, #3, and #4. An unevenly distributed exhaust structure is exemplified. Alternatively, like the intake passage 30, the exhaust pipe 43 may be positioned at the center in the width direction of the exhaust manifold 42, or the independent exhaust pipes 44_#1 to #4 may have the same flow path length. good.

(4) In the above embodiment, the EGR prediction unit 64 predicts the internal EGR amount using the polynomial model shown in FIG. Without using the polynomial model, for example, table data that associates the operating conditions with the internal EGR amount may be stored in advance in the storage unit 68, and the table data may be read out when predicting the internal EGR amount.

1 engine body 2 cylinder #1, #4 cylinder (first cylinder/end cylinder)
#2, #3 cylinders (second cylinder/central cylinder)
9 intake port 10 exhaust port 11 intake valve 12 exhaust valve 13a intake VVT (internal EGR mechanism)
14a exhaust VVT (internal EGR mechanism)
15 injector (fuel injection valve)
30 intake passage (intake route)
303 third passage (upstream intake passage)
36 surge tank 36R internal space (flow path space)
37 independent intake passage 60 processor (engine combustion control device)
61 fuel injection control unit (control unit)

Claims (5)

multiple cylinders and
a surge tank arranged in an intake path to the plurality of cylinders;
an independent intake passage connecting the surge tank and each intake port of the plurality of cylinders;
an intake valve and an exhaust valve that respectively open and close each intake port and each exhaust port of the plurality of cylinders;
a fuel injection valve disposed for each of the plurality of cylinders and supplying fuel to each cylinder;
an internal EGR mechanism that achieves internal EGR by setting a valve overlap that opens both the intake valve and the exhaust valve;
A control unit that controls the fuel injection amount of each of the fuel injection valves according to the operating state of the engine ,
An engine capable of executing a first combustion mode in which partial compression ignition combustion combining SI combustion and CI combustion is performed in an environment substantially matching the stoichiometric air-fuel ratio, and a second combustion mode in which the combustion is performed in a lean environment. A combustion control device for
The independent intake passage is set to have a length that allows the intake air blown out from the intake port to reach the surge tank during the period of the valve overlap,
The control unit is
In both the first combustion mode and the second combustion mode, the target fuel injection amount for each cylinder determined according to the operating state of the engine is set to the re-intake amount of intake air from the intake port by internal EGR in each cylinder. Correction based on the re-suction correction amount set for each cylinder according to
Further, in the first combustion mode, the re-suction correction amount is corrected according to the engine speed , and the re-suction correction amount is corrected to decrease as the engine speed increases. An engine combustion control device characterized by: In the engine combustion control device according to claim 1 ,
A combustion control device for an engine, wherein the control unit corrects the re-intake correction amount so as to decrease as the intake pressure increases. In the engine combustion control device according to claim 1 or 2 ,
Among the plurality of cylinders, a first cylinder having a predetermined amount of re-intake of intake air from the intake port by the internal EGR, and a second cylinder having a larger re-intake amount than the first cylinder. exists and
The control unit corrects the target fuel injection amount so that the air-fuel ratio of the first cylinder is leaner and the air-fuel ratio of the second cylinder is richer. . In the engine combustion control device according to claim 1 or 2 ,
The engine is a four-cylinder engine in which four cylinders are arranged in a row,
The independent intake passages are arranged in a row in the direction in which the cylinders are arranged, and the surge tank has a flow passage space elongated in the direction in which the cylinders are arranged,
The surge tank is connected to an upstream intake passage that introduces intake air into a central region in the longitudinal direction of the surge tank,
The control unit adjusts the air-fuel ratio of the end cylinders located at both ends of the four cylinders arranged in a line to a lean side, and the air-fuel ratio of the two central cylinders sandwiched between the end cylinders. A combustion control device for an engine that corrects the target fuel injection amount so as to move toward the rich side. In the engine combustion control device according to any one of claims 1 to 3 ,
The control unit includes, as elements, at least a valve overlap amount at which both the intake valve and the exhaust valve of the cylinder are open, an intake/exhaust differential pressure, a closing timing of the exhaust valve, and an engine speed. An engine combustion control device that sets the reintake correction amount based on the reintake amount calculated using a polynomial model of .

Images