JP2021008857A - Combustion control device of engine - Google Patents

Abstract

To suppress a variation of torque between cylinders without causing the deterioration of fuel economy and an increase of NOx, and to prevent the deterioration of combustion stability resulting from a combustion variation in a specified cylinder, in an engine which can perform partial compression ignition combustion.SOLUTION: A combustion control unit 61 has a correction unit 67 for correcting a fuel injection amount. The correction unit 67 primarily corrects the fuel injection amount to each of the cylinders 2a to 2d by an injector 15 on the basis of a deviation between target combustion timing which is acquired by predicting combustion gravity center timing and actual combustion timing in which the combustion gravity center timing is acquired on the basis of an actual combustion state so that periods from ignition timing up to the combustion gravity center timing at which fuel at a 50 mass% ratio is combusted are aligned between and among the cylinders (step S01). Also, when a combustion variation becomes larger than a prescribed threshold in a specified cylinder (YES in step S03), the correction unit 67 secondarily corrects the fuel injection amount which is set by a primary correction with respect to the specified cylinder to an increase side (step S04).SELECTED DRAWING: Figure 7

Description

The present invention relates to a combustion control device for an engine capable of partial compression ignition combustion in which a part of the air-fuel mixture is SI-combusted by spark ignition and the other air-fuel mixture is CI-combusted by self-ignition.

As a combustion mode of a gasoline engine, a part of the air-fuel mixture is forcibly burned by flame propagation (SI combustion) triggered by spark ignition using a spark plug, and the other air-fuel mixture is burned by self-ignition (CI combustion). Partial compression ignition combustion (SPCCI combustion) is known. In Patent Document 1, the optimum time when the combustion mass ratio becomes 50%, that is, the time when 50% of the fuel injected into the combustion chamber burns in the engine in which partial compression ignition combustion is performed is set as a target value. The technology to be defined is disclosed. In Patent Document 1, by controlling the ignition timing so that the main combustion timing becomes the target value, compression self-ignition combustion is stably realized without being affected by the operating state of the engine.

Japanese Patent No. 3873580

By the way, in a multi-cylinder engine, fuel injection control is performed for each cylinder in order to suppress variations in torque between cylinders. Normally, the fuel injection amount is controlled so that the air-fuel ratio (A / F) of each cylinder is the same. However, although the control of aligning the A / F can be expected to improve the variation in torque between cylinders, it may lead to deterioration of fuel efficiency and an increase in NOx depending on the operating condition of the engine. In particular, in an engine that performs partial compression ignition combustion, this tendency is observed in the operating region where the air-fuel ratio is larger than the stoichiometric air-fuel ratio (λ> 1).

In view of this problem, in the above operating region, the present inventors can control the fuel injection amount so that the gas-fuel ratio (G / F) is uniform rather than the A / F between cylinders. It was found to be effective from the viewpoint of improving fuel efficiency and reducing NOx. However, even if the control to align the G / F is executed, combustion fluctuation (torque fluctuation) may occur in a specific cylinder, and as a result, there is a new problem that the stability of compression self-ignition combustion is lowered. Encountered.

An object of the present invention is to suppress fluctuations in torque between cylinders without deteriorating fuel efficiency or increasing NOx in an engine capable of partial compression ignition combustion, and combustion caused by combustion fluctuations in a specific cylinder. It is an object of the present invention to provide an engine combustion control device capable of preventing deterioration of stability.

The combustion control device of the engine according to one aspect of the present invention includes a plurality of cylinders, an injector for injecting fuel into the cylinders, and an ignition plug for generating sparks in the cylinders, and mixes fuel and air in the cylinders. A combustion control device for an engine in which a part of the air is SI-combusted by spark ignition and then the remaining air-fuel mixture in the cylinder is CI-combusted by self-ignition. The combustion control device is the ignition plug of each cylinder and the ignition plug. A combustion control unit that controls the combustion state of each cylinder by controlling the operation of each injector is provided, and the combustion control unit of the ignition plug so that a predetermined specific combustion time appears at a predetermined time. Target combustion obtained by predicting the predetermined mass combustion timing so that the ignition control unit that controls the ignition timing and the period from the ignition timing to the predetermined mass combustion timing at which a predetermined mass ratio of fuel burns are aligned in each cylinder. A first correction unit that primarily corrects the fuel injection amount to each cylinder by the injector based on the deviation between the timing and the actual combustion timing obtained by determining the predetermined mass combustion timing based on the actual combustion state, and a specific cylinder. When the combustion fluctuation becomes larger than a predetermined threshold value, a second correction unit that secondarily corrects the fuel injection amount set by the primary correction for the specific cylinder to the increase side is included. ..

According to the combustion control device of this engine, it is possible to substantially align the G / F of each cylinder by the primary correction of the first correction unit. Unlike the A / F of each cylinder, which can be estimated relatively accurately based on the sensor output, it is difficult to accurately grasp the G / F of each cylinder. However, the present inventors have a predetermined mass combustion timing in which a predetermined mass ratio of fuel is burned from the ignition timing by the spark plug, for example, a fuel having a mass ratio of 50% of the fuel supplied to each cylinder during one combustion cycle. It was found that there is a correlation between the period until the combustion center of combustion and the G / F. That is, it was found that if the period from the ignition timing to the predetermined mass combustion timing in each cylinder is aligned, it is substantially equivalent to aligning the G / F of each cylinder. Therefore, by executing the primary correction, it is possible to suppress the variation in torque between cylinders while suppressing the deterioration of fuel efficiency and the increase in NOx.

On the other hand, even if the G / F of each cylinder is controlled to be aligned, combustion fluctuation may occur in a specific cylinder for some reason. When combustion fluctuation occurs, the torque generated per cycle fluctuates with time. This combustion fluctuation can be suppressed by increasing the fuel injection amount to the cylinder in which the combustion fluctuation occurs. Therefore, the fuel injection amount to a specific cylinder in which a combustion fluctuation larger than a predetermined threshold value is corrected by the second correction unit to the increase side by the secondary correction, thereby causing the combustion fluctuation of the specific cylinder. It can be suppressed. As a result, combustion stability can be ensured.

The combustion control device includes a plurality of in-cylinder pressure sensors for detecting the pressure in each cylinder, and the first correction unit grasps the actual combustion timing based on the detection value of the in-cylinder pressure sensor. can do.

According to this combustion control device, the first correction unit simply grasps the time when the predetermined mass combustion time is reached (actual combustion time) in the actual combustion based on the in-cylinder pressure sequentially detected by the in-cylinder pressure sensor. be able to. However, an error may intervene in specifying the actual combustion timing based on the detection result of the in-cylinder pressure sensor. For example, in a low-load operating region where the rise of the combustion pressure (in-cylinder pressure) is small, it is difficult to specify the combustion center of gravity timing depending on the combustion pressure change characteristic, so an error is likely to occur. In this case, the fuel injection amount was intended to be primarily corrected so that the G / F was aligned between the cylinders, but the accurate primary correction was not completed because the actual combustion timing was not specified accurately, and the combustion fluctuation was caused in a specific cylinder. Can be invited. Therefore, when the in-cylinder pressure sensor is used, the embodiment of the present invention in which the fuel injection amount set by the primary correction is secondarily corrected to the increase side and the combustion fluctuation is eliminated becomes particularly useful.

In the above combustion control device, the first correction unit performs the primary correction so that the actual combustion time approaches the target combustion time, and the actual combustion time is retarded from the target combustion time. In this case, the primary correction is performed so as to increase the fuel injection amount, and the second correction unit sets an offset value for retarding the actual combustion timing as the secondary correction. It is desirable to increase the fuel injection amount for the specific cylinder in the primary correction executed thereafter.

According to this combustion control device, the secondary correction is executed by utilizing the control mode of the primary correction that the fuel injection amount is increased when the actual combustion time is retarded from the target combustion time. Can be made to. That is, by setting the offset value, it is possible to create a state in which the actual combustion timing is virtually retarded. Then, in the primary correction executed thereafter, the fuel injection amount is increased so as to bring the virtually retarded actual combustion timing closer to the target combustion timing, and as a result, the secondary correction is executed. It will be. Therefore, since the secondary correction is performed using the control system for the primary correction without preparing a separate control system for the secondary correction, the control system can be simplified. it can.

In the upper combustion control device, the first correction unit detects the target combustion period, which is the period from the ignition timing to the target combustion timing and is obtained based on a preset combustion model, and the in-cylinder pressure sensor. It is desirable to first correct the fuel injection amount of each cylinder based on the deviation from the actual combustion period, which is the period from the ignition timing to the actual combustion timing, which is obtained based on the in-cylinder pressure.

According to this combustion control device, the target combustion period can be set reliably and quickly based on the combustion model. Then, based on the deviation between the target combustion period and the actual combustion period in each cylinder, so to speak, the estimated and actual difference, the period from the ignition timing of each cylinder to the predetermined mass timing can be quickly adjusted by feedback control or the like by the primary correction. It will be possible.

In this case, the first correction unit sets an average value of individual differences, which is a deviation between the target combustion period and the actual combustion period for each cylinder, as a correction target value, and sets the correction target value and the individual difference. It is more desirable to first correct the fuel injection amount of each cylinder so that the deviation becomes zero.

According to this combustion control device, as a result of adjusting the actual combustion period to the target combustion period in each cylinder, the actual combustion period between the cylinders is not aligned, but the correction target value which is the average value of the individual differences is set. By combining the individual differences of each cylinder, the actual combustion period between the cylinders can be aligned. Therefore, it is possible to align the actual combustion period between the cylinders in the actual combustion period that does not necessarily match the target combustion period, and it is possible to increase the degree of freedom in aligning the actual combustion period.

In the above combustion control device, the engine is subjected to lean operation in which the air-fuel ratio, which is the ratio of air and fuel in the cylinder, is higher than the stoichiometric air-fuel ratio in at least a part of the operating region. It is desirable that the combustion control unit executes the primary correction and the secondary correction in the partial operating region.

In partial compression ignition combustion, which is a combination of SI combustion and CI combustion, it is particularly important to align the G / F rather than the A / F in order to suppress the variation in torque between cylinders in the situation where lean combustion is performed. It is effective. Therefore, according to the above-mentioned combustion control device, in the operating region where the lean operation is executed, the torque variation between cylinders is suppressed without deterioration of fuel efficiency and increase of NOx, and combustion in a specific cylinder is performed. It is possible to prevent deterioration of combustion stability due to fluctuations.

In the above combustion control device, the engine includes an exhaust manifold that guides exhaust gas discharged from each of the plurality of cylinders, and the exhaust manifold is located at a position closest to one of the plurality of cylinders. The first correction unit is provided between the one cylinder and the remaining other cylinders when the internal EGR is executed in the plurality of cylinders. It is desirable to carry out the primary correction so as to suppress the variation in the generated torque between the cylinders due to the difference in the internal EGR rate.

When the internal EGR is executed, in an engine having an exhaust manifold in which the gathering part is unevenly distributed at a position close to one cylinder as described above, the EGR of the one cylinder is performed even if the overlap period of the supply and exhaust strokes is the same. The amount of gas is larger than that of other cylinders, and the combustion speed of the one cylinder tends to be slower. That is, the actual combustion period in the one cylinder tends to be longer than that in the same period of the other cylinders. However, according to the above-mentioned combustion control device, the variation in the generated torque between the cylinders due to the difference in the internal EGR ratio between the one cylinder and the remaining other cylinders is suppressed. That is, it is possible to suppress the variation in the generated torque between the cylinders caused by the form of the exhaust manifold.

According to the present invention, in an engine capable of partial compression ignition combustion, it is possible to suppress the variation in torque between cylinders without deteriorating fuel efficiency or increasing NOx, and combustion caused by combustion fluctuation in a specific cylinder. It is possible to provide an engine combustion control device that can prevent deterioration of stability.

FIG. 1 is a system diagram schematically showing an overall configuration of a compression ignition type engine to which the combustion control device of the engine according to the present invention is applied. FIG. 2 is a plan view showing the engine body and a part of the exhaust passage. FIG. 3 is a block diagram showing an engine control system. FIG. 4 is an operation map in which the operating area of the engine is divided according to the difference in the control of the combustion mode. FIG. 5 is a graph showing a waveform of the heat generation rate during SPCCI combustion (partial compression ignition combustion). FIG. 6 is a graph showing an example of combustion fluctuation for each cylinder. FIG. 7 is a flowchart showing a basic operation of combustion control according to the present invention. FIG. 8 is a graph for explaining a mode of primary correction of the fuel injection amount. FIG. 9 is a graph for explaining a mode of secondary correction of the fuel injection amount. FIG. 10 is a flowchart showing combustion control by the ECU during SPCCI combustion. It is a flowchart (subroutine) which shows the detail of the feedback correction amount calculation in the flowchart of FIG. FIG. 12 is a diagram showing an example of first fuel correction data (λ> 1) for correcting the target fuel injection amount in the primary correction, where (a) is # 1 cylinder and (b) is # 2. Cylinders, (c) shows correction data for # 3 cylinder, and (d) shows correction data for # 4 cylinder. FIG. 13 is a diagram showing an example of second fuel correction data for correcting the target fuel injection amount.

Hereinafter, embodiments of the present invention will be described with reference to the drawings. In the present embodiment, an example in which the combustion control device of the engine according to the present invention is applied to a compression ignition type engine will be described. The compression ignition type engine can execute partial compression ignition combustion in which a part of the fuel and air mixture in the cylinder is SI-combusted by spark ignition and then the remaining mixture in the cylinder is CI-combusted by self-ignition. Engine.

[Engine system]
First, an engine system including the compression ignition type engine will be described. FIG. 1 is a diagram showing an overall configuration of an engine system according to the present embodiment. The engine system is a 4-cycle 4-cylinder gasoline direct injection engine mounted on a vehicle as a power source for traveling, and includes an engine body 1, an intake passage 30 through which intake air introduced into the engine body 1 flows, and an engine body. It includes an exhaust passage 40 through which the exhaust gas discharged from No. 1 flows, and an EGR device 50 that returns a 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 a piston 5. The cylinder block 3 has a cylinder liner that forms four cylinders 2 (a plurality of cylinders / see # 1 to # 4 cylinders 2a to 2d in FIG. 2) arranged in a row. Note that FIG. 1 shows only one cylinder 2. The cylinder head 4 is attached to the upper surface of the cylinder block 3 and closes the upper opening of the cylinder 2. The piston 5 is housed in each cylinder 2 so as to be slidable back and forth, and is connected to the crankshaft 7 via a connecting rod 8. The crankshaft 7 rotates around its central axis in response to the reciprocating motion of the piston 5.

A combustion chamber 6 is formed above the piston 5. 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. The air-fuel mixture of the supplied fuel and air is burned in the combustion chamber 6, and the piston 5 pushed down by the expansion force due to the combustion reciprocates in the vertical direction. The geometric compression ratio of the cylinder 2, that is, the ratio of the volume of the combustion chamber 6 when the piston 5 is at the top dead center and the volume of the combustion chamber 6 when the piston 5 is at the bottom dead center is determined by SPCCI combustion described later. A high compression ratio of 13 or more and 30 or less, preferably 14 or more and 18 or less is set so as to be suitable for (partial compression ignition combustion).

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 the cooling water flowing inside the cylinder block 3 and the cylinder head 4, that is, the engine water temperature.

The cylinder head 4 is formed with an intake port 9 and an exhaust port 10 that communicate with the combustion chamber 6. The bottom surface of the cylinder head 4 is the ceiling surface of the combustion chamber 6. An intake side opening which is a downstream end of the intake port 9 and an exhaust side opening which is an upstream end of the exhaust port 10 are formed on the ceiling surface of the combustion chamber. The cylinder head 4 is assembled with 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. Although not shown, the valve type of the engine body 1 is a 4-valve type of 2 intake valves x 2 exhaust valves, and two intake ports 9 and two exhaust ports 10 are provided for each cylinder 2. At the same time, 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 including a camshaft and an exhaust side valve mechanism 14. The intake valve 11 and the exhaust valve 12 are opened and closed by the valve operating mechanisms 13 and 14 in conjunction with the rotation of the crankshaft 7. The intake side valve mechanism 13 has a built-in intake VVT 13a capable of changing at least the opening time of the intake valve 11. Similarly, the exhaust side valve mechanism 14 includes an exhaust VVT 14a capable of changing at least the closing time 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, an internal EGR that leaves a high-temperature burned gas in the combustion chamber 6 is realized. Further, it is possible to adjust the internal EGR amount (residual amount of burnt gas) by adjusting the valve overlap amount, which is the period during which the valve overlap is performed.

In the cylinder head 4, an injector 15 and a spark plug 16 are further arranged for each of the four cylinders 2. The injector 15 injects fuel into the cylinder 2. As the injector 15, a multi-injector type injector having a plurality of injection holes at its tip and capable of injecting fuel radially from these injection holes can be used. The injector 15 is arranged so 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 includes a tip electrode portion that generates a spark in the cylinder 2. The spark plug 16 is arranged at a position slightly offset from the injector 15 on the intake side, and the tip electrode portion is arranged at a position facing 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 a plurality of in-cylinder pressure sensors SN3 corresponding to each cylinder 2 as sensing elements. The in-cylinder pressure sensor SN3 detects the pressure (in-cylinder pressure) in each cylinder 2 (combustion chamber 6). Based on the detected value of the in-cylinder pressure sensor SN3, the actual combustion time described later is grasped.

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. The air (fresh air) taken in from the upstream end of the intake passage 30 is introduced into the combustion chamber 6 through the intake passage 30 and the intake port 9. An air cleaner 31, a throttle valve 32, a supercharger 33, an electromagnetic clutch 34, an intercooler 35, and a surge tank 36 are arranged in the intake passage 30 in this order from the upstream side.

The air cleaner 31 removes foreign matter in the intake air to purify the intake air. The throttle valve 32 opens and closes the intake passage 30 in conjunction with the depression operation 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 the electromagnetic clutch 34 switches between engaging and releasing the engagement with the engine body 1. When the electromagnetic clutch 34 is engaged, the driving force is transmitted from the engine body 1 to the supercharger 33, and supercharging is performed by the supercharger 33. The intercooler 35 cools the intake air compressed by the supercharger 33. The surge tank 36 is a tank that is arranged in the intake path to the four cylinders 2 and provides a space for evenly distributing the intake air to the four cylinders 2.

A swirl valve 18 is arranged in one of the two intake ports 9 provided corresponding to each cylinder 2. By adjusting the opening degree of the swirl valve 18, the strength of the swirl flow that swirls around the central axis of the combustion chamber 6 can be adjusted. The intake port 9 of the present embodiment is a tumble port capable of forming a tumble flow. Therefore, the swirl flow formed when the swirl valve 18 is closed becomes an oblique swirl flow mixed with the tumble flow.

In each part of the intake passage 30, an air flow sensor SN4 for detecting the flow rate of the intake air, first and second intake air temperature sensors SN5 and SN7 for detecting the temperature of the intake air, and first and second suction for detecting the pressure of the intake air. Atmospheric pressure sensors SN6 and SN8 are provided. The air flow sensor SN4 and the first intake air temperature sensor SN5 are arranged in a portion of the intake passage 30 between the air cleaner 31 and the throttle valve 32, and detect the flow rate and temperature of the intake air passing through the portion. 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 from the connection port of the EGR passage 51 described later), and the intake air passing through the portion is provided. Detect pressure. The second intake air temperature sensor SN7 is provided in a portion of the intake passage 30 between the supercharger 33 and the intercooler 35, and detects the temperature of the intake air passing through the portion. The second intake pressure sensor SN8 is provided in the surge tank 36 and detects the pressure of the 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 the 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, which will be described later, to each other. The bypass passage 38 is provided with a bypass valve 39 capable of opening and closing 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. The burnt 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 catalyst 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 a particulate matter (PM) contained in the exhaust gas. It has a built-in GPF (gasoline particulate filter) 41b for collecting. In addition, another catalyst converter containing an appropriate catalyst such as a three-way catalyst or a NOx catalyst may be added to the downstream side of the catalyst converter 41.

A linear O ₂ sensor SN10 for detecting the concentration of oxygen contained in the exhaust gas is arranged in a portion of the exhaust passage 40 on the upstream side of the catalytic converter 41. The linear O ₂ sensor SN10 is a type of sensor whose output value changes linearly according to the density of oxygen concentration. It is possible to estimate the air-fuel ratio of the air-fuel mixture based on the output value of the linear O ₂ sensor SN10.

The EGR device 50 includes an EGR passage 51 that connects the exhaust passage 40 and the intake passage 30, an EGR cooler 52 and an EGR valve 53 provided in the EGR passage 51. The EGR passage 51 connects a portion of the exhaust passage 40 on the downstream side of the catalytic converter 41 and a portion of the intake passage 30 between the throttle valve 32 and the supercharger 33 to each other. The EGR cooler 52 cools the exhaust gas (external EGR) that is returned from the exhaust passage 40 to the intake passage 30 through the EGR passage 51 by heat exchange. The EGR valve 53 is provided 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 SN9 for detecting the difference between the pressure on the upstream side and the pressure on the downstream side of the EGR valve 53.

The accelerator 17 is provided with an accelerator sensor SN11 that detects the accelerator opening degree. The accelerator sensor SN11 is a sensor that detects the degree of pedal depression of the accelerator 17, and is also a sensor that detects the driver's acceleration / deceleration intention.

[Structural features of the exhaust passage]
Next, the structural features of the exhaust passage 40 described above will be described. FIG. 2 is a plan view showing the engine body 1 and a part of the exhaust passage 40. Four cylinders 2 (# 1 cylinder 2a, # 2 cylinder 2b, # 3 cylinder 2c and # 4 cylinder 2d) are arranged in a row in the engine body 1. The exhaust passage 40 includes an exhaust manifold 42 connected to the exhaust port 10 of the engine body 1. The exhaust manifold 42 is attached to the side surface of the cylinder head 4 and guides (collects) the exhaust gas discharged from each of the exhaust ports 10 of the # 1 to # 4 cylinders 2. The exhaust gas collected in the exhaust manifold 42 is guided downstream by an exhaust pipe connected to the catalytic converter 41.

The exhaust manifold 42 is a gathering unit for merging the first to fourth independent exhaust pipes 43a to 43d corresponding to each of the # 1 to # 4 cylinders 2a to 2d and the exhaust gas discharged from these independent exhaust pipes 43a to 43d. It is equipped with 44. The first independent exhaust pipe 43a is in the exhaust port 10 of the # 1 cylinder 2a, the second independent exhaust pipe 43b is in the exhaust port 10 of the # 2 cylinder 2b, and the third independent exhaust pipe 43c is in the exhaust port 10 of the # 3 cylinder 2c. The fourth independent exhaust pipe 43d communicates with the exhaust port 10 of the # 4 cylinder 2d. The collecting portion 44 is a portion that gathers the downstream end portions of the first to fourth independent exhaust pipes 43a to 43d. The catalytic converter 41 is arranged on the downstream side of the gathering portion 44. The burnt gas generated in the combustion chambers 6 of the cylinders 2a to 2d is collected in the collecting portion 44 through the independent exhaust pipes 43a to 43d of the exhaust manifold 42, and is introduced into the catalytic converter 41 through the collecting portion 44.

The collecting portion 44 is provided at a position closest to the # 1 cylinder 2a (one cylinder) located on the rightmost side in FIG. 2 in the arrangement direction of the # 1 to # 4 cylinders 2a to 2d. Therefore, the lengths of the independent exhaust pipes 43a to 43b from each exhaust port 10 to the collecting portion 44 are not uniform lengths, but are biased. Specifically, the length of the first independent exhaust pipe 43a corresponding to the # 1 cylinder 2a on the right end side is the shortest, and the second to fourth independent exhaust pipes 43b to the other # 2 to # 4 cylinders 2b to 2d The length of 43d is # 2, # 3, # 4, which becomes longer toward the left side. Such a difference in the lengths of the independent exhaust pipes 43a to 43b can affect the internal EGR rate.

[Control system]
FIG. 3 is a block diagram showing a control system of the engine system. The engine system of this embodiment is collectively controlled by the ECU 60 (engine combustion control device). The ECU 60 is a microprocessor composed of a CPU, ROM, RAM, and the like.

Detection signals from various sensors are input to the ECU 60. The ECU 60 includes the crank angle sensor SN1, water temperature sensor SN2, in-cylinder pressure sensor SN3, airflow sensor SN4, first and second intake temperature sensors SN5, SN7, first and second intake pressure sensors SN6, SN8, and differential pressure sensor. SN9, are connected linear _{O 2} sensor SN10 and an accelerator sensor SN11 and electrically. Information detected by these sensors SN1 to SN11, that is, crank angle, engine rotation speed, engine water temperature, in-cylinder pressure, intake flow rate, intake temperature, intake pressure, exhaust gas oxygen concentration, front-rear differential pressure of EGR valve 53. , Information such as the accelerator opening degree is sequentially input to the ECU 60.

The ECU 60 controls each part of the engine while executing various determinations and calculations based on the input information from each of the above sensors. That is, the ECU 60 is electrically connected to the intake VVT 13a, the exhaust VVT 14a, the injector 15, the spark plug 16, the swirl valve 18, the throttle valve 32, the electromagnetic clutch 34, the bypass valve 39, the EGR valve 53, and the like. A control signal is output to each of these devices based on the results and the like.

The ECU 60 operates so as to functionally include the combustion control unit 61 and the storage unit 68 by executing a predetermined program. The combustion control unit 61 controls the combustion state of the cylinders 2a to 2d by controlling the operations of the injectors 15 and the spark plugs 16 of the # 1 to # 4 cylinders 2a to 2d. The storage unit 68 stores various data, set values, calculation formulas, and the like. In the present embodiment, a polynomial combustion model or the like for obtaining a target combustion period, which will be described later, is stored in the storage unit 68.

When a predetermined program is executed, the combustion control unit 61 includes an operation state determination unit 62, an ignition control unit 63, an injection control unit 64, an EGR control unit 65, a combustion fluctuation determination unit 66, and a correction unit 67 (first correction). It operates so as to functionally include a unit and a second correction unit).

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 accelerator opening degree information detected by the accelerator sensor SN11. This determination result is used to determine which region of the predetermined driving map the current driving region is.

The ignition control unit 63 controls the ignition timing of the spark plug 16. The ignition timing is controlled so that a predetermined specific combustion timing appears at a predetermined timing. In the present embodiment, for example, the ignition timing is controlled so that the timing of switching from SI combustion to CI combustion (inflection point X in FIG. 5) becomes the target crank angle (θci).

The injection control unit 64 controls the injection amount and injection pattern of the fuel injected from the injector 15 according to various conditions. The injection control unit 64 sets the target fuel injection amount and the injection pattern of each cylinder 2 according to the accelerator depression amount (engine operating state) detected by the accelerator sensor SN11. Further, the injection control unit 64 causes the injector 15 to execute the fuel injection operation with the set target fuel injection amount and injection pattern. When the target fuel injection amount is corrected by the correction unit 67 described later, the injection control unit 64 causes the injector 15 to perform fuel injection at the corrected fuel injection amount.

The EGR control unit 65 executes the external EGR by controlling the opening degree of the EGR valve 53 of the EGR device 50. Further, the EGR control unit 65 executes the internal EGR by appropriately setting the valve overlap period of the intake valve 11 and the exhaust valve 12.

The combustion fluctuation determination unit 66 monitors the combustion fluctuation in each cylinder 2 with reference to the detected value of the in-cylinder pressure sensor SN3. The in-cylinder pressure sensor SN3 detects the in-cylinder pressure of the cylinder 2. Therefore, by monitoring the in-cylinder pressure, it is possible to grasp the maximum pressure, the combustion period, etc. in each combustion cycle, and further, it is possible to grasp the fluctuation for each of these cycles, that is, the combustion fluctuation. Further, the combustion fluctuation determination unit 66 determines whether or not the combustion fluctuation is larger than a predetermined threshold value. This determination result is given to the correction unit 67. The threshold value is stored in the storage unit 68 in advance.

The correction unit 67 has each of the # 1 to # 4 cylinders 2a to 2d so that the period from the ignition timing of the air-fuel mixture by the spark plug 16 to the predetermined mass combustion timing at which the fuel having a predetermined mass ratio burns is aligned. The fuel injection amount to each cylinder 2a to 2d by the injector 15 is primarily corrected. The predetermined mass combustion timing can be set to various combustion timings, but in the present embodiment, a fuel having a mass ratio of 50% of the fuels supplied to the cylinders 2a to 2d during one combustion cycle. The time of the center of gravity of combustion when the fuel burns. The correction unit 67 primaryly corrects the fuel injection amount based on the deviation between the target combustion timing obtained by predicting the combustion center of gravity time and the actual combustion time obtained by obtaining the combustion center of gravity time based on the actual combustion state. Specifically, the primary correction is a feedback correction that increases or decreases the target fuel injection amount set by the injection control unit 64 for each cylinder 2a to 2d so that the deviation is eliminated as much as possible.

The correction unit 67 further secondarily corrects the fuel injection amount when the combustion fluctuation in the specific cylinder 2 becomes larger than a predetermined threshold value. Even if the combustion centers of gravity of the cylinders 2a to 2d are aligned by the above primary correction, combustion fluctuations that cannot be overlooked may occur for some reason. In this case, the correction unit 67 further secondary-corrects the fuel injection amount set by the primary correction for the specific cylinder having a large combustion fluctuation to the increase side. The secondary correction here is, for example, if A / F = 30 after the primary correction, the correction is such that the fuel injection amount is increased so that A / F = 29. The primary correction and the secondary correction executed by the correction unit 67 will be described in more detail later.

[Driving map]
FIG. 4 is an operation map used when the engine (engine body 1) is warm, and is a diagram in which the operation area of the engine is divided according to the difference in the control of the combustion mode. In the following description, a high (low) engine load is equivalent to a high (low) engine required torque. When the engine is in a warm state, the operating region of the engine is roughly divided into three operating regions A1, A2, and A3 according to the difference in combustion form. These operating areas A1 to A3 are referred to as a first operating area A1, a second operating area A2, and a third operating area A3, respectively.

The third operating region A3 is a high-speed region where the engine speed is high. The first operating region A1 is a low / medium speed / low load region excluding a part of the high load side from the region on the low speed side of the third operating region A3. The second operating region A2 is a residual region other than the first and third operating regions A1 and A3, that is, a low / medium speed / high load region. Hereinafter, the combustion modes and the like selected in each operating region will be described in order.

<First and second operating areas>
In the low / medium speed / low load first operating region A1 and the low / medium speed / high load second operating region A2, partial compression ignition combustion that combines SI combustion and CI combustion (hereinafter referred to as SPCCI combustion). ) Is executed. SI combustion is a combustion mode in which the air-fuel mixture is ignited by sparks generated from the spark plug 16 and the air-fuel mixture is forcibly burned by flame propagation that expands the combustion region from the ignition point to the surroundings. CI combustion is a combustion form in which the air-fuel mixture is burned by self-ignition in an environment where the temperature and pressure are sufficiently increased by compression of the piston 5. SPCCI combustion, which is a combination of these SI combustion and CI combustion, is the SI combustion of a part of the air-fuel mixture in the combustion chamber 6 by spark ignition performed in an environment just before the air-fuel mixture self-ignites. This is a combustion mode in which the other air-fuel mixture in the combustion chamber 6 is CI-combusted by self-ignition later (due to further increase in temperature and pressure accompanying SI combustion). "SPCCI" is an abbreviation for "Spark Controlled Compression Ignition".

FIG. 5 is a graph showing a combustion waveform when SPCCI combustion is performed as described above, that is, a change in the heat generation rate (J / deg) with respect to the crank angle. As shown in the figure, in SPCCI combustion, heat generation due to SI combustion and heat generation due to CI combustion are continuously generated in this order. At this time, due to the property that the combustion speed of CI combustion is faster, the rise of heat generation becomes steeper during CI combustion than during SI combustion. Therefore, the waveform of the heat generation rate in SPCCI combustion has an inflection point X that appears at the timing of switching from SI combustion to CI combustion (θci described later).

Here, in SPCCI combustion, which is a combination of SI combustion and CI combustion, the ratio of SI combustion and CI combustion is controlled according to the operating conditions. As an example of control, each part of the engine is controlled so that the SI rate, which is the ratio of the heat generated by SI combustion to the total heat generated by SPCCI combustion, becomes an appropriate value.

The SI rate can be defined as follows. In FIG. 5, the crank angle θci (predetermined time) corresponding to the inflection point X (specific combustion time) at which the combustion form switches from SI combustion to CI combustion is defined as the start time of CI combustion. In this case, the amount of heat generated by SI combustion corresponds to the area R1 of the waveform of the heat generation rate on the advance side of the θci, and the amount of heat generated by CI combustion is the heat located on the retard side of the θci. It can be considered to correspond to the area R2 of the waveform of the occurrence rate. Then, the SI rate can be defined as R1 / (R1 + R2) by using each of these areas R1 and R2.

In the first and second operating regions A1 and A2 where SPCCI combustion is performed, each part of the engine is controlled so that the above-mentioned SI rate and θci match a predetermined target value (crank angle). That is, in the first and second operating regions A1 and A2, the target SI rate, which is the target value of the SI rate, and the target θci, which is the target value of θci, are set for each of various conditions in which the engine load and the engine speed are different. Be done. Then, various control items are controlled so as to be a feasible combination of the target SI rate and the target θci. The control items include the ignition timing of the spark plug 16 controlled by the ignition control unit 63 (FIG. 3), the fuel injection amount / injection timing from the injector 15 controlled by the injection control unit 64, and the external control by the EGR control unit 65. The EGR rate, the internal EGR rate, and the like. The external EGR ratio is the weight ratio of the external EGR gas (exhaust gas recirculated to the combustion chamber 6 through the EGR passage 51) in the total gas in the combustion chamber 6. The internal EGR ratio is the weight ratio of the internal EGR gas (burnt gas remaining in the combustion chamber 6 due to the internal EGR) to the total gas in the combustion chamber 6.

For example, the ignition timing and the fuel injection amount / injection timing are determined by a predetermined map in consideration of the target SI rate and the target θci. That is, in the map, the ignition timing and the fuel injection amount / injection timing suitable for achieving the target SI rate and the target θci are described for each condition of the engine load and the engine speed. The map is stored in the storage unit 68. The ignition control unit 63 and the injection control unit 64 control the spark plug 16 and the injector 15 according to the ignition timing and the fuel injection amount / injection timing stored in the map.

On the other hand, the external EGR rate and the internal EGR rate are determined by calculation using a predetermined model formula. That is, the EGR control unit 65 uses a predetermined model formula for the in-cylinder temperature (target in-cylinder temperature) required at the time of spark ignition in order to realize the target SI rate and the target θci for each combustion cycle. To calculate. Then, the EGR control unit 65 determines the opening degree of the EGR valve 53 and the valve timings of the intake valve 11 and the exhaust valve based on the calculated target cylinder temperature.

In the first and second operating regions A1 and A2, the throttle valve 32 is controlled as follows in addition to the above-mentioned ignition timing and injection amount / injection timing control. That is, in the first operating region A1, basically, more air than the amount of air equivalent to the stoichiometric air-fuel ratio is introduced into the combustion chamber 6 through the intake passage 30, that is, the air in the combustion chamber 6 (new). The opening degree of the throttle valve 32 so that the air-fuel ratio (A / F), which is the weight ratio of q) and fuel, becomes larger than the stoichiometric air-fuel ratio (14.7) (the excess air ratio λ> 1). Is set. On the other hand, in the second operating region A2, the opening degree at which the amount of air equivalent to the stoichiometric air-fuel ratio is introduced into the combustion chamber 6, that is, the air-fuel ratio substantially matches the stoichiometric air-fuel ratio (λ≈1). , The opening degree of the throttle valve 32 is set.

<Third operating area>
In the third operating region A3, which has a higher rotation speed than the first and second operating regions A1 and A2, normal SI combustion is executed. For example, fuel is injected from the injector 15 for a predetermined period that overlaps at least a part of the intake stroke, and spark ignition by the spark plug 16 is executed in the latter half of the compression stroke. Then, SI combustion is started triggered by this spark ignition, and all of the air-fuel mixture in the combustion chamber 6 is burned by flame propagation.

In the third operating region A3, the throttle valve 32 has an opening degree such that an air amount equivalent to the stoichiometric air-fuel ratio or a smaller amount of air is introduced into the combustion chamber 6, that is, the air-fuel ratio in the combustion chamber 6 is theoretical. The opening is set so that the air-fuel ratio or a value slightly richer than this (λ ≦ 1) is obtained.

In the first to third operating areas A1 to A3, the supercharger 33 is turned off in the inner region of the supercharging line T shown in FIG. 4, and the supercharger 33 is turned off in the outer region of the supercharging line T. Is turned on. In the inner region of the supercharging line T, that is, on the low speed side of the first operating region A1, the electromagnetic clutch 34 is released, the connection between the supercharger 33 and the engine body 1 is released, and the bypass valve 39 is fully opened. As a result, supercharging by the supercharger 33 is stopped. On the other hand, in the outer region of the supercharging line T, that is, on the high-speed side of the first operating region A1, the electromagnetic clutch 34 is engaged and the supercharger 33 and the engine body 1 are connected to each other, so that the supercharger 33 supercharges The salary is done. At this time, the pressure (supercharging pressure) in the surge tank 36 detected by the second intake pressure sensor SN8 matches the target pressure predetermined for each engine operating condition (conditions such as rotation speed and load). As described above, the opening degree of the bypass valve 39 is controlled.

[Primary correction to suppress torque variation between cylinders]
Subsequently, the primary correction of the fuel injection amount executed by the correction unit 67 (first correction unit) will be described in detail. In the engine body 1, it is important to suppress variations in torque generated by the cylinders 2a to 2d in order to ensure stable running performance. Therefore, it is necessary to control the combustion of each cylinder 2a to 2d according to the operating state (rotational speed / load) of the engine. In this case, if the combustion conditions of the cylinders 2a to 2d are the same, the fuel injection amount / injection timing and the like may be controlled so that the air-fuel ratios (A / F) of the cylinders 2a to 2d of the engine are the same. ..

However, in reality, the combustion conditions of the cylinders 2a to 2d are not always the same, and control with uniform A / F may lead to deterioration of fuel efficiency and an increase in NOx depending on the operating state of the engine. In particular, this tendency is observed in the first operating region A1 (SPCCI_λ> 1) in which the air-fuel ratio becomes larger than the stoichiometric air-fuel ratio in SPCCI combustion. One of the factors is the influence of variation in the internal EGR rate (amount of internal EGR). The internal EGR ratio of each cylinder 2a to 2d is uniformly determined by calculation using a predetermined model formula as described above. Then, the valve overlap period of the intake valve 11 and the exhaust valve 12 in each cylinder 2a to 2d is controlled so as to realize the determined internal EGR rate.

However, there is a discrepancy between the internal EGR ratio based on the model formula and the actual internal EGR ratio, and the degree of the discrepancy also varies between the cylinders 2a to 2d. In the present embodiment, the EGR rate of the # 1 cylinder 2a tends to be higher than the EGR rate of each of the other cylinders 2b to 2d. This is because, as shown in FIG. 2, the exhaust manifold 42 of the present embodiment has a structural feature that the # 1 cylinder 2a is ubiquitously close to the gathering portion 44 with respect to the # 2 to # 4 cylinders 2b to 2d. Has. Such structural features can contribute to the compactification of the engine system and the like. However, due to the structural features of the exhaust manifold 42, compared to # 2 to # 4 cylinders 2b to 2d (other cylinders), the burned gas is # 1 cylinder 2a (1) when the internal EGR is performed. It becomes easy to be pulled back to the cylinder). Such a variation in the return amount of the burned gas is expressed as a difference in the internal EGR rate between the cylinders 2a and 2d, which in turn causes a variation in the generated torque between the cylinders 2a and 2d when the internal EGR is executed. To do.

Therefore, the inventors of the present application have focused on the gas-fuel ratio (G / F), which is the ratio of the mass of gas including air and burned gas to the mass of fuel. That is, attention was paid to controlling the fuel injection amount so that the G / Fs of the cylinders 2a to 2d are aligned instead of the A / F. Then, in the first operating region A1, it is tested that by controlling the G / F alignment, it is possible to suppress the variation in torque between the cylinders 2a to 2d while suppressing the deterioration of fuel efficiency and the increase in NOx. Confirmed.

However, unlike the A / F that can be estimated relatively accurately by the sensor output of the airflow sensor SN4 or the like, it is difficult to accurately grasp the G / F of each cylinder 2a to 2d by the sensor output. In view of this point, the inventors of the present application have found that if the combustion centers of gravity of each cylinder 2a to 2d are aligned, the G / F will also be aligned. FIG. 5 shows the combustion center of gravity timing (θmfb50; predetermined mass combustion timing). The combustion center of gravity period is the period when 50% of the fuel supplied to the cylinder during one combustion cycle burns. There may be a correlation between the period θt (hereinafter referred to as 50% combustion period θt) from the ignition start timing (ignition timing θig) by the spark plug 16 to the combustion center of gravity timing (θmfb50) and the G / F. found. Specifically, it was found that if the 50% combustion periods θt in each cylinder 2a to 2d are aligned, the G / F of each cylinder can be substantially aligned. FIG. 5 shows a time (crank angle) at which θci (the start time of CI combustion) corresponding to the inflection point X at which SI combustion is switched to CI combustion and the time of the center of gravity of combustion θmfb50 are different (crank angle). In FIG. 5, for convenience of illustration, θci and θmfb50 are separated from each other, but both periods are substantially the same.

In view of the above findings, the correction unit 67 is set by the injection control unit 64 according to the current operating state so that the 50% combustion period θt from the ignition timing θig to the combustion center of gravity timing θmfb50 is aligned in each cylinder 2a to 2d. The fuel injection amount by each injector 15 is firstly corrected. In this primary correction, the fuel injection is performed so that the deviation between the target combustion time obtained by predicting the 50% combustion period θt and the actual combustion time obtained by obtaining the 50% combustion period θt based on the actual fuel condition is eliminated. Feedback correct the amount.

Specifically, the correction unit 67 as the first correction unit that performs this primary correction corrects the fuel injection amount of each cylinder 2a to 2d based on the deviation between the target combustion period and the actual combustion period. The target combustion period is a period from the ignition timing θig to the target combustion timing, which is a period obtained by a preset combustion model. The combustion model here is a model in which an ideal combustion mode is predetermined for each operating state under predetermined conditions. The actual combustion period is a period from the ignition timing θig to the actual combustion timing, and is a period obtained based on the in-cylinder pressure detected by the in-cylinder pressure sensor SN3.

[Secondary correction to suppress combustion fluctuation]
When a specific cylinder 2 in which the combustion fluctuation is larger than a predetermined threshold value occurs, the correction unit 67 as the second correction unit makes a secondary correction so as to increase the fuel injection amount of the specific cylinder 2. When the combustion fluctuation occurs, the torque generated per cycle of the specific cylinder 2 fluctuates with time, which may affect the running performance. Such combustion fluctuations can be suppressed by increasing the fuel injection amount to the cylinder in which the combustion fluctuations occur. That is, in a situation where lean combustion of SPCCI is performed and combustion fluctuation in which the in-cylinder pressure is not stable occurs, increasing the fuel injection amount is effective for stabilizing combustion.

More details will be given with reference to FIG. FIG. 6 is a graph showing an example of combustion fluctuation for each cylinder. Here, an example of combustion fluctuation for each cycle of # 1 to # 4 cylinders 2a to 2d is shown. The combustion fluctuation on the vertical axis is an index indicated by the ratio between the illustrated mean effective pressure (IMEP) based on the in-cylinder pressure detected by the in-cylinder pressure sensor SN3 and its standard deviation. Even if the 50% combustion period θt is controlled to be uniform in each cylinder 2a to 2d by the above primary correction, combustion fluctuation may occur for some reason. FIG. 6 shows an example in which a large combustion fluctuation exceeding a predetermined threshold value Th occurs in the # 4 cylinder 2d. For the other # 1 to # 3 cylinders 2a to 2c, the combustion fluctuation is within the threshold Th.

An error in the detection value of the in-cylinder pressure detected by the in-cylinder pressure sensor SN3 can be cited as a cause of such combustion fluctuation occurring in the specific cylinder 2d. As described above, in the first correction, the actual combustion time is grasped based on the in-cylinder pressure of each cylinder 2a to 2d sequentially detected by the in-cylinder pressure sensor SN3. However, an error may intervene in specifying the actual combustion time based on the detection result of the in-cylinder pressure sensor SN3. For example, in a low rotation and low load operating region where the rise of the combustion pressure (in-cylinder pressure) is small, it is difficult to specify the combustion center of gravity timing θmfb50 based on the in-cylinder pressure change characteristic, so that an error is likely to occur. In particular, in the present embodiment, lean combustion of SPCCI is applied to the first operating region A1 (partial operating region) of low load and low rotation, the fuel injection amount is inherently small, and the rise of the in-cylinder pressure change is small. It can also be said that it is easy to induce an error.

For this reason, although the fuel injection amount is intended to be corrected so that the G / F is aligned between the cylinders in the primary correction, the accurate primary correction may not be completed because the actual combustion time is not specified accurately. In the example of FIG. 6, as a result of the primary correction not being accurate for the # 4 cylinder 2d, it can be inferred that the combustion environment for each cycle is not stable and the combustion fluctuation exceeds the threshold Th.

In this case, the correction unit 67 further secondarily corrects the fuel injection amount set by the primary correction for the # 4 cylinder 2d having a large combustion fluctuation to the increase side. In general, increasing the fuel injection amount for a cylinder in which combustion fluctuations occur can lead to stabilization of combustion. The correction unit 67 adjusts so that the total fuel injection amounts of all the cylinders 2a to 2d do not fluctuate so that the torque generated by the entire engine does not become excessive when the secondary correction is performed. Specifically, the correction unit 67 proportionally divides and reduces the fuel injection amount of the other # 1 to # 3 cylinders 2a to 2c by the amount of the increased fuel injection amount of the # 4 cylinder 2d.

[Basic Embodiment of Combustion Control and Its Variations]
Subsequently, the basic operation of correcting the fuel injection amount in the combustion control according to the present invention will be described with reference to the flowchart shown in FIG. 7. The correction unit 67 acquires the target fuel injection amount set by the injection control unit 64 according to the depression amount of the accelerator 17 detected by the accelerator sensor SN11. Then, the correction unit 67 sets each cylinder 2a to 2d so that the 50% combustion period θt (actual combustion time) in actual combustion is aligned with the intention of aligning the G / F of each cylinder 2a to 2d. The target fuel injection amount is primarily corrected (step S01).

FIG. 8 is a graph for explaining a mode of primary correction of the fuel injection amount. The heat generation rate waveform H shown by the solid line in FIG. 8 is a waveform equivalent to the heat generation rate waveform in SPCCI combustion shown in FIG. 5 above. It is assumed that the 50% combustion period θt from the ignition timing θig to the combustion center of gravity timing θmfb50 of the heat generation rate waveform H is the target combustion period obtained by the above-mentioned combustion model. The correction unit 67 obtains a deviation between the 50% combustion period θt and the actual combustion period in each of the cylinders 2a to 2d, and feedback-corrects the fuel injection amount so that the deviation is eliminated.

The heat generation rate waveform H1 shown by the dotted line in FIG. 6 shows a combustion state in which the actual combustion period from the ignition timing θig to the combustion center of gravity timing is longer than the target 50% combustion period θt. That is, in the heat generation rate waveform H1, the 50% combustion time θdel obtained based on the actual combustion state appears on the retard side of the combustion center of gravity time θmfb50. When such a discrepancy occurs, the fuel injection amount is corrected so that the 50% combustion time θdel approaches the target combustion center of gravity time θmfb50. Specifically, the correction unit 67 performs the primary correction so as to increase the fuel injection amount for the cylinder 2 in which combustion such as the heat generation rate waveform H1 is occurring. That is, the progress of combustion is promoted by increasing the fuel injection amount, and the 50% combustion time θdel appears on the advance side from the current state.

On the other hand, the heat generation rate waveform H2 shown by the alternate long and short dash line in FIG. 6 shows a combustion state in which the actual combustion period from the ignition timing θig to the combustion center of gravity timing is shorter than the target 50% combustion period θt. In the heat generation rate waveform H2, the 50% combustion time θadv obtained based on the actual combustion state appears on the advance side of the combustion center of gravity time θmfb50. In this case, the correction unit 67 makes the primary correction so as to reduce the fuel injection amount for the cylinder 2 in which combustion such as the heat generation rate waveform H2 is occurring. That is, the progress of combustion is slowed down by the reduction of the fuel injection amount, the 50% combustion time θadv appears on the retard side from the current state, and the 50% combustion period θt approaches.

In the above embodiment, as a result of adjusting the actual combustion period of the # 1 to # 4 cylinders 2a to 2d to the 50% combustion time θt which is the target combustion period, the actual combustion period between the cylinders 2a to 2d is aligned. is there. Instead of this, in another embodiment, the individual difference which is the deviation between the target combustion period and the actual combustion period may be obtained for each of the cylinders 2a to 2d, and the average value of these individual differences may be set as the correction target value. .. Then, the fuel injection amount of each cylinder 2a to 2d is firstly corrected so that the deviation between the correction target value and the individual difference becomes zero. That is, instead of forcibly adjusting the actual combustion period between the cylinders 2a to 2d to the predicted 50% combustion time θt, the individual of each cylinder 2a to 2d is set to the correction target value which is the average value of the individual differences. It is a mode in which the actual combustion period is made uniform by adjusting the differences. This aspect will be illustrated later (FIG. 11).

Next, the combustion fluctuation determination unit 66 confirms the combustion fluctuation in the # 1 to # 4 cylinders 2a to 2d with reference to the detected value of the in-cylinder pressure sensor SN3 (step S02). This is a process for grasping the combustion fluctuation characteristics for each of the cylinders 2a to 2d as illustrated in FIG. Subsequently, the combustion fluctuation determination unit 66 determines whether or not the specific cylinders in which the combustion fluctuations exceeding the predetermined threshold Th are exist in the # 1 to # 4 cylinders 2a to 2d (step S03). .. If the specific cylinder does not exist (NO in step S03), the process returns to step S01 and the process is repeated.

On the other hand, when the specific cylinder is present (YES in step S03), the correction unit 67 increases the fuel injection amount by a predetermined amount for the specific cylinder in which the combustion fluctuation exceeding the threshold Th is occurring. The next correction is executed (step S04). After that, the process returns to step S01 and the process is repeated. In the next process, if the combustion fluctuation of the specific cylinder falls within a predetermined threshold value Th due to the increase in the fuel injection amount, a NO determination is made in step S03, and the fuel injection amount is not further increased. On the other hand, even with the increase in the fuel injection amount, if the combustion fluctuation of the specific cylinder still exceeds the predetermined threshold Th, the fuel injection amount is further increased stepwise.

The flow of FIG. 7 shows an example in which the primary correction and the secondary correction are used together. Instead of this, when the specific cylinder occurs (YES in step S03), giving priority to stabilizing the combustion fluctuation, as shown by the dotted line in FIG. 7, the process returns to step S02. You may repeat it. That is, the primary correction of step S01 may be temporarily stopped.

In addition, in order to make the torque uniform between the cylinders, it is conceivable to monitor the combustion fluctuation preferentially with respect to other factors and perform control to stabilize the combustion fluctuation. However, it takes a considerable amount of time to evaluate the combustion fluctuation. That is, unless a considerable number of detected values of the in-cylinder pressure sensor SN3 are acquired and statistical processing is performed, the combustion fluctuation cannot be accurately evaluated. On the other hand, the timing of the center of gravity of combustion can be obtained immediately. Therefore, as shown in FIG. 7, it is desirable to control the torque between the cylinders to be uniform based on the primary correction, and to compensate for the combustion fluctuation by the secondary correction when it exceeds the permissible range.

Further, the primary correction and the secondary correction may be achieved by an independent control system. That is, the secondary correction may be separated from the primary correction and the fuel injection amount may be simply increased stepwise. However, it is desirable to simplify the control system by substantially incorporating the fuel injection amount increasing process by the secondary correction into the routine of the primary correction. This point will be described with reference to FIG. FIG. 9 is a graph for explaining a desirable mode of secondary correction of the fuel injection amount.

As shown in FIG. 8, the correction unit 67 performs the primary correction so that the actual combustion time (θdel or θadv) approaches the target combustion time (combustion center of gravity time θmfb50), and the actual combustion time is When the angle is retarded from the target combustion timing (heat generation rate waveform H1), the primary correction is performed so as to increase the fuel injection amount. The secondary correction can be executed by utilizing the correction mode of the primary correction. Specifically, the correction unit 67 sets an offset value for retarding the actual combustion timing of the specific cylinder 2 whose combustion fluctuation exceeds the threshold value Th as the secondary correction. As a result, it is possible to increase the fuel injection amount for the specific cylinder in the primary correction executed thereafter.

With reference to FIG. 9, it is assumed that the heat generation rate waveform H0 shown by the solid line in the figure is observed in the specific cylinder 2 in which the large combustion fluctuation occurs. As described in FIG. 5, the heat generation rate waveform H0 has a 50% combustion period θt (actual combustion timing) from the ignition timing θig to the combustion center of gravity timing θmfb50. In this case, the correction unit 67 adds the offset value α to the combustion center of gravity timing θmfb50 on the retard side to create a virtual heat generation rate waveform Hα. The virtual heat generation rate waveform Hα is a waveform having a combustion center of gravity time θmfb50 + α at a crank angle retarded by an offset value α from the combustion center of gravity time θmfb50 of the actual heat generation rate waveform H0.

When the virtual heat generation rate waveform Hα as described above is created for the specific cylinder 2, in the primary correction executed thereafter, the correction unit 67 sets the virtually retarded combustion center of gravity timing θmfb50 + α to the actual combustion. It is treated as a time, and the actual combustion center of gravity time θmfb50 is treated as the target combustion time. Therefore, the correction unit 67 performs a primary correction for increasing the fuel injection amount of the specific cylinder 2 in order to bring the virtual combustion center of gravity time θmfb50 + α closer to the actual combustion center of gravity time θmfb50. That is, since the fuel injection amount of the specific cylinder 2 is increased so as to advance the combustion center of gravity timing by the offset value α, it is possible to obtain a result substantially equal to the execution of the secondary correction. According to this embodiment, the secondary correction is performed using the control system for the primary correction without preparing a separate control system for the secondary correction, so that the control system can be used. It can be simplified.

[Specific combustion control flow]
Subsequently, a more specific flow of combustion control according to the present invention will be described. FIG. 10 is a flowchart showing control by the ECU 60 in the first and second operation areas A1 and A2 in the operation map of FIG. 4 during operation by SPCCI combustion. FIG. 11 is a flowchart showing the process (subroutine) of step S9 of the flowchart of FIG.

When the control shown in this flowchart starts, the ECU 60 reads various signals from the sensors SN1 to SN11 shown in FIG. 3 and other sensors, and acquires information on the operating state of the engine body 1 and environmental information that affects the combustion conditions. (Step S1). Subsequently, the combustion control unit 61 specifies the engine load specified from the engine rotation speed detected by the crank angle sensor SN1, the detected value of the accelerator sensor SN11 (accelerator opening), the detected value of the airflow sensor SN4 (intake flow rate), and the like. Based on the above, the target θci, which is the target value of the crank angle θci (FIG. 5) corresponding to the turning point X (specific combustion time), is determined (step S2). Then, the ignition control unit 63 and the injection control unit 64 determine the target ignition timing and the target fuel injection amount / target injection timing suitable for realizing the target θci (step S3). The target θci, the target fuel injection amount / target injection timing, and the target ignition timing are determined based on a predetermined map or the like.

Next, the operating state determination unit 62 is lean (λ>) in which the current current operating point is in the first operating region A1 based on the engine speed and the engine load, that is, the air-fuel ratio is larger than the theoretical air-fuel ratio. 1) It is determined whether or not the SPCCI combustion is in the operating region in the state of (step S4). When SPCCI_λ> 1 combustion is executed (YES in step S4), the correction unit 67 functioning as the first correction unit is based on the first fuel correction data for SPCCI_λ> 1 stored in advance in the storage unit 68. Then, the target fuel injection amount determined in step S3 is corrected (step S5).

The first fuel correction data for SPCCI_λ> 1 is the internal EGR rate of each cylinder 2a to 2d obtained based on a predetermined model formula, the tendency of deviation from the actual internal EGR rate, and the actual internal EGR between cylinders. This is data in which the fuel correction amount is determined so that the deviation and the variation are corrected based on the tendency of the rate variation. Specifically, the relationship between the internal EGR rate (theoretical value) based on the model formula and the fuel correction amount is defined. In addition to the above-mentioned data for the first operating region A1, the data for the second operating region A2 and the data for SPCCI_λ = 1 are set in the first fuel correction data. In step S11, which will be described later, the target fuel injection amount is corrected based on the first fuel correction data for SPCCI_λ = 1.

12 (a) to 12 (d) are graphs showing an example of first fuel correction data of # 1 to # 4 cylinders 2a to 2d in the operating region of SPCCI_λ> 1. The correction amount of the target fuel injection amount shown in this graph is a unique correction amount determined by the characteristic structure (FIG. 2) of the exhaust manifold 42 adopted in the present embodiment. 12A shows the first fuel correction data of # 1 cylinder 2a, (b) shows the # 2 cylinder 2b, (c) shows the # 3 cylinder 2c, and (d) shows the first fuel correction data of the # 4 cylinder 2d. The storage unit 68 stores the table data of the correction amount corresponding to these graphs.

For the # 1 cylinder 2a closest to the collecting portion 44, the amount of re-intake of burnt gas from the exhaust port 10 is relatively large. On the other hand, for # 2 to # 4 cylinders 2b to 2d, the amount of re-inhaled burned gas is relatively small. In this case, if the # 1 to # 4 cylinders 2a to 2d are made to inject fuel according to the target fuel injection amount, the amount of gas containing air becomes excessive in the # 1 cylinder 2a, and the G / F becomes lean, and conversely. In the # 2 to # 4 cylinders 2b to 2d, the amount of gas becomes insufficient, the G / F becomes rich, and the G / F varies. To correct this variation, correct the target fuel injection amount so that the G / F is toward the rich side for the # 1 cylinder 2a and the G / F is toward the lean side for the # 2 to # 4 cylinders 2b to 2d. Just do it.

Specifically, as shown in FIG. 12A, a correction amount (+) for increasing the target fuel injection amount is set for the # 1 cylinder 2. As a result, the # 1 cylinder 2 whose G / F fluctuates toward the lean side due to an excessive amount of re-inhalation of burnt gas can be brought closer to the target G / F. Further, the degree of increase in the target fuel injection amount is set so as to increase as the internal EGR rate increases. This is because as the amount of internal EGR increases, the amount of burned gas returning from the exhaust passage 40 increases, and the G / F variation between the cylinders 2a and 2d tends to increase. On the other hand, as shown in FIGS. 12 (b), 12 (c), and (d), for # 2, # 3, and # 4 cylinders 2b to 2d, the correction amount (-) for reducing the target fuel injection amount is It is set, and the degree of weight loss is set so as to increase as the internal EGR rate increases. By making such a correction, it is possible to bring the # 2, # 3, and # 4 cylinders 2b to 2d, which are rich in G / F due to insufficient return of the burnt gas, closer to the target G / F.

Specifically, in step S5, the correction unit 67 applies a polynomial model stored in the storage unit 68 to calculate a provisional value of the internal EGR rate of each cylinder 2a to 2d. The polynomial model is a polynomial model having valve overlap amount, valve closing timing of the exhaust valve 12, intake / exhaust differential pressure, and engine speed as elements. Subsequently, the correction unit 67 calculates the predicted value of the internal EGR rate by multiplying the calculated provisional value of the internal EGR rate by the coefficient determined by the exhaust gas temperature and the coefficient determined by the atmospheric pressure. Then, the correction unit 67 determines the fuel correction amount of each cylinder 2a to 2d based on the obtained internal EGR rate and the first fuel correction data (FIG. 12), and determines the fuel correction amount of each cylinder 2a to 2d based on the fuel correction amount. Correct the target fuel injection amount.

Next, the correction unit 67 further corrects the corrected target fuel injection amount in step S5 based on the feedback correction amount obtained in the previous combustion cycle (step S6). As a result, the final target fuel injection amount is determined.

When the final target fuel injection amount is determined, the correction unit 67 uses the above-mentioned combustion model to determine the combustion center of gravity timing θmfb50 (target combustion timing) based on the target fuel injection amount and the current operating state (rotational speed / load). ) Is specified. Further, the correction unit 67 calculates the predicted 50% combustion period θt0 (target combustion period), which is the period from the ignition timing θig in the combustion model to the specified combustion center of gravity timing θmfb50, for each cylinder 2a to 2d (step S7). ).

Next, the injection control unit 64 injects the fuel of the final target fuel injection amount determined in step S6 by the injector 15 at the target injection timing determined in step S3. Further, the ignition control unit 63 causes the spark plug 16 to perform an ignition operation at the ignition timing determined in step S3 (step S8). As a result, SPCCI_λ> 1 combustion is executed.

After that, the correction unit 67 calculates a feedback correction amount for aligning the 50% combustion period θt between the cylinders 2a and 2d, that is, eliminating the variation in the generated torque between the cylinders 2a and 2d (step S9). The subroutine of the calculation process of the feedback correction amount will be described with reference to the flowchart shown in FIG.

When the subroutine starts, the correction unit 67 obtains the combustion center of gravity time θmfb50 (actual combustion time) in the SPCCI combustion actually performed in step S8. Further, the correction unit 67 obtains the actual 50% combustion period θt1 (actual combustion period), which is the period from the ignition timing θig to the combustion center of gravity timing θmfb50, by calculation (step S21). Specifically, the correction unit 67 calculates the amount of heat generated by combustion for each crank angle based on the waveform of the in-cylinder pressure detected by the in-cylinder pressure sensor SN3 during the period of SPCCI combustion, and for each crank angle. Based on the heat generation amount data, the combustion center of gravity time θmfb50 in which 50% by mass of the fuel is burned is calculated. Then, the actual 50% combustion period θt1 is derived by obtaining the period from the ignition timing θig determined in step S3 to the combustion center of gravity timing θmfb50.

Next, the correction unit 67 sets the individual difference Δθ, which is the deviation between the predicted 50% combustion period θt0 obtained in step S7 of FIG. 10 and the actual 50% combustion period θt1 calculated in the previous step S21, from the cylinders 2a to 2d. The calculation is performed every time (step S22). Further, the correction unit 67 obtains an average value of the individual differences Δθ of each cylinder 2a to 2d, and sets this average value as the correction target value θtv (step S23).

After that, the correction unit 67 calculates the deviation between the correction target value θtv and the individual difference Δθ of each cylinder 2a to 2d (step S24). Basically, the fuel injection amount is firstly corrected so that the deviation between the correction target value θtv and the individual difference Δθ goes to zero. Subsequently, the correction unit 67 determines the degree of reflection of the individual difference Δθ with respect to the target fuel injection amount based on the calculation result of the deviation and the predetermined second fuel correction data (step S25). The degree of reflection is a parameter indicating whether or not the target fuel injection amount is to be corrected and to what extent the correction is made, and the correction coefficient and the like are examples of the degree of reflection.

FIG. 13 shows an example of the second fuel correction data. The second fuel correction data defines the relationship between the deviation between the target value θtv and the individual difference Δθ (the value calculated in step S24) and the degree of reflection. In the present embodiment, the above relationship is generally set so that the degree of reflection becomes relatively large as the deviation between the target value θtv and the individual difference Δθ becomes relatively large. However, when the deviation between the target value θtv and the individual difference Δθ exceeds a specific value, the degree of reflection is maintained at a constant value (maximum value).

When the reflection degree for each of the cylinders 2a to 2d is determined, the correction unit 67 calculates a specific fuel correction amount (feedback correction amount) for the target fuel injection amount for each cylinder 2a to 2d based on the reflection degree. Step S26). This feedback correction amount is calculated, for example, by substituting a numerical value indicating the degree of reflection obtained in step S25 into a predetermined model formula.

After that, the process proceeds to the second correction process. The combustion fluctuation determination unit 66 determines whether or not there is a specific cylinder 2 in which the combustion fluctuation in each cylinder 2a to 2d is larger than the predetermined threshold value Th (FIG. 6) based on the monitoring result. Determine (step S27). When the specific cylinder 2 does not exist (NO in step S27), the processing of the subroutine ends.

On the other hand, when the specific cylinder 2 exists (YES in step S27), the correction unit 67 functioning as the second correction unit sets an offset value for retarding the actual 50% combustion period θt1 of the specific cylinder 2. To do. That is, the process of adding the offset value α previously described with reference to FIG. 7 to the combustion center of gravity timing θmfb50 obtained in step S21 is executed (step S28). After that, the offset value α is added to the calculation of the actual 50% combustion period θt1 of the specific cylinder 2. As described above, this process substantially secondarily corrects the fuel injection amount of the specific cylinder 2 toward the increase side. Note that the addition of the offset value α is maintained even if the combustion fluctuation of the specific cylinder 2 falls below the threshold value Th in the routine after the next time by setting the offset value α. This is because if the addition of the offset value α is removed, there is a high possibility that the combustion fluctuation of the specific cylinder 2 will worsen again. However, once the operation region of SPCCI_λ> 1 combustion is deviated, the addition of the offset value α may be reset.

When the subroutine of FIG. 11 is completed, the correction unit 67 updates the feedback correction amount calculated in step S26 to the storage unit 68 (step S10). Of course, when the offset value α is set in the specific cylinder 2 in step S28, the offset value α is also stored. After that, the process returns to step S1.

On the other hand, in step S4 of the flowchart of FIG. 10, when it is determined that the current operation point is not the first operation area A1 (NO in step S4), that is, the operation state determination unit 62 is the second operation area A2. If the determination is made, the correction unit 67 shifts the process to step S11. In step S11, the correction unit 67 corrects the target fuel injection amount determined in step S3 based on the first fuel correction data for SPCCI_λ = 1 stored in advance in the storage unit 68.

As for the first fuel correction data for SPCCI_λ = 1, the relationship between the internal EGR rate (theoretical value) based on the model formula and the combustion correction amount was determined as in the case of the first fuel correction data for SPCCI_λ> 1. It is a thing. The first fuel correction data for SPCCI_λ = 1 is exclusively the internal EGR rate and the combustion correction so as to correct the deviation between the internal EGR rate of each cylinder 2a to 2d based on the model formula and the actual internal EGR rate. The relationship with quantity is defined.

Next, the injection control unit 64 injects the fuel of the final target fuel injection amount determined in step S11 by the injector 15 at the target injection timing determined in step S3. Further, the ignition control unit 63 causes the spark plug 16 to perform an ignition operation at the ignition timing determined in step S3 (step S12). As a result, SPCCI_λ = 1 combustion is executed. After that, the ECU 60 returns the process to step S1.

[Action effect]
According to the combustion control device of the engine according to the present embodiment described above, each cylinder 2a is subjected to the primary correction of the fuel injection amount for aligning the actual 50% combustion period θt1 of each cylinder 2a to 2d, which is executed by the correction unit 67. It is possible to substantially align the G / F of ~ 2d. As a result, it is possible to suppress the variation in torque between the cylinders 2a to 2d while suppressing the deterioration of fuel efficiency and the increase in NOx. On the other hand, even if the G / F of each cylinder 2a to 2d is controlled to be aligned, combustion fluctuation exceeding a predetermined threshold value Th may occur in a specific cylinder for some reason. In this case, the correction unit 67 performs secondary correction to increase the fuel injection amount to a specific cylinder in which a large combustion fluctuation occurs, and suppresses the combustion fluctuation. Therefore, combustion stability can be ensured.

Further, the correction unit 67 sets an offset value α that retards the actual 50% combustion period θt1 as the secondary correction, so that the fuel injection amount for the specific cylinder is set in the primary correction executed thereafter. Increase the amount. That is, by setting the offset value α, it is possible to create a state in which the actual 50% combustion period θt1 is virtually retarded. As a result, in the primary correction routine executed thereafter, the fuel injection amount is increased in order to bring the virtually retarded actual 50% combustion period θt1 closer to the target combustion time, and as a result, the secondary correction is performed. Will be executed. Therefore, since the secondary correction is performed using the control system for the primary correction without preparing a separate control system for the secondary correction, the control system can be simplified. it can.

Further, in the embodiment shown in FIG. 11, the average value of the individual difference Δθ is set as the target value θtv, and the feedback correction amount is determined based on the deviation between the target value θtv and the individual difference Δθ of each cylinder 2a to 2d. Will be done. That is, since the feedback correction amount is obtained so that the individual difference Δθ of each cylinder 2a to 2d approaches the target value θtv, the feedback correction amount of each cylinder 2a to 2d is not constrained by the predicted 50% combustion period θt1 based on the combustion model. The 50% combustion period θt can be made uniform.

1 Engine body 2 Cylinders 2a, 2b, 2c, 2d # 1, # 2, # 3, # 4 cylinders 15 Injectors 16 Spark plugs 42 Exhaust manifold 60 ECU (engine combustion control device)
61 Combustion control unit 63 Ignition control unit 67 Correction unit (first correction unit, second correction unit)
θig Ignition timing θmfb50 Combustion center of gravity timing (predetermined mass combustion timing)
Crank angle corresponding to θci inflection point X (predetermined time)
SN3 In-cylinder pressure sensor

Claims (7)

A plurality of cylinders, an injector that injects fuel into the cylinders, and a spark plug that generates sparks in the cylinders are provided, and a part of the fuel-air mixture in the cylinders is SI-combusted by spark ignition and then the cylinders. It is a combustion control device of an engine that CI burns the remaining air-fuel mixture by self-ignition.
The combustion control device includes a combustion control unit that controls the combustion state of each cylinder by controlling the operation of the spark plug and the injector of each cylinder.
The combustion control unit
An ignition control unit that controls the ignition timing of the spark plug so that a predetermined specific combustion timing appears at a predetermined timing.
The target combustion timing obtained by predicting the predetermined mass combustion timing and the predetermined mass combustion timing are actually set so that the period from the ignition timing to the predetermined mass combustion timing in which the fuel having a predetermined mass ratio burns is the same for each cylinder. Based on the deviation from the actual combustion timing obtained based on the combustion state of, the first correction unit that primary corrects the fuel injection amount to each cylinder by the injector, and
An engine including a second correction unit that secondarily corrects the fuel injection amount set by the primary correction for the specific cylinder to the increase side when the combustion fluctuation becomes larger than a predetermined threshold value in the specific cylinder. Combustion control device. In the combustion control device for the engine according to claim 1.
Equipped with multiple in-cylinder pressure sensors that detect the pressure in each cylinder,
The first correction unit is an engine combustion control device that grasps the actual combustion timing based on the detection value of the in-cylinder pressure sensor. In the combustion control device for the engine according to claim 1.
The first correction unit performs the primary correction so that the actual combustion time approaches the target combustion time, and when the actual combustion time is retarded from the target combustion time, the fuel injection amount is adjusted. The primary correction is performed so as to increase the amount.
The second correction unit sets an offset value for retarding the actual combustion timing as the secondary correction, thereby increasing the fuel injection amount for the specific cylinder in the primary correction executed thereafter. Engine combustion control device. In the combustion control device for the engine according to claim 2.
The first correction unit is a period from the ignition timing to the target combustion timing, which is obtained based on the target combustion period obtained based on a preset combustion model and the in-cylinder pressure detected by the in-cylinder pressure sensor. , An engine combustion control device that primarily corrects the fuel injection amount of each cylinder based on the deviation from the actual combustion period, which is the period from the ignition timing to the actual combustion timing. In the combustion control device for the engine according to claim 4.
The first correction unit sets an average value of individual differences, which are deviations between the target combustion period and the actual combustion period for each cylinder, as a correction target value, and the deviation between the correction target value and the individual difference is zero. An engine combustion control device that primarily corrects the fuel injection amount of each cylinder so as to go toward. The engine combustion control device according to any one of claims 1 to 5.
The engine is subjected to lean operation in which the air-fuel ratio, which is the ratio of air to fuel in the cylinder, is higher than the stoichiometric air-fuel ratio in at least a part of the operating region.
The combustion control unit is an engine combustion control device that executes the primary correction and the secondary correction in the partial operating region. In the combustion control device for the engine according to claim 6.
The engine includes an exhaust manifold that guides exhaust gas discharged from each of the plurality of cylinders, and the exhaust manifold is discharged from each cylinder at a position closest to one of the plurality of cylinders. Equipped with a gathering part where exhaust gas collects
When the internal EGR is executed in the plurality of cylinders, the first correction unit determines the variation in torque generated between the cylinders due to the difference in the internal EGR ratio between the one cylinder and the remaining other cylinders. An engine combustion control device that performs the primary correction so as to suppress it.

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