JP7255368B2 - engine controller - Google Patents

Description

The present invention relates to a device for controlling a compression ignition engine capable of self-igniting at least part of an air-fuel mixture in a cylinder.

2. Description of the Related Art Engines installed in vehicles and the like are required to suppress combustion noise to a predetermined level or less. On the other hand, for example, in Patent Document 1, in an engine equipped with a spark plug and configured so that the air-fuel mixture self-ignites triggered by ignition by the spark plug, the ignition timing is retarded when the combustion noise increases. , thereby reducing combustion noise. Specifically, the engine of Patent Document 1 is provided with a pressure sensor that detects the pressure in the cylinder, and it is determined whether or not the pressure in the cylinder detected by the pressure sensor is higher than a predetermined value. When the pressure in the cylinder is higher than a predetermined value, it is estimated that the combustion noise has exceeded a desired level, and the timing for igniting the air-fuel mixture in the cylinder is retarded.

JP 2019-39383 A

If the ignition timing is retarded, the start and peak times of combustion can sometimes be made later in the expansion stroke, and in this case, an increase in combustion noise can be suppressed. However, in an engine configured so that at least a portion of the air-fuel mixture self-ignites, there are cases where retarding the ignition timing cannot suppress the rapid rise in cylinder pressure and the increase in combustion noise, depending on the state of the air-fuel mixture. be. Specifically, before the mixture is ignited, or when the mixture is already in a self-ignitable state at the time the mixture is ignited, the mixture is mixed regardless of ignition timing. The air will start to self-ignite. Therefore, in this case, even if the ignition timing is retarded, it is not possible to suppress the sudden increase in pressure in the cylinder and the increase in combustion noise. Further, in an engine that self-ignites an air-fuel mixture without using a spark plug, it is of course impossible to suppress an increase in combustion noise by retarding the ignition timing.

SUMMARY OF THE INVENTION It is an object of the present invention to provide an engine control system capable of reliably preventing an increase in combustion noise.

In order to solve the above-mentioned problems, the present invention includes fuel injection means for injecting fuel into a cylinder, and self-ignites at least part of a mixture of fuel and air injected from the fuel injection means. A device for controlling a compression ignition engine that can be combusted by, a determination means for determining the presence or absence of abnormal combustion in which the progress of combustion of an air-fuel mixture is excessively fast, a first fuel injection, and the first injection control means for causing the fuel injection means to perform a second fuel injection whose injection timing is later than the first fuel injection, wherein the injection control means determines by the determination means that the abnormal combustion does not occur. the fuel injection means is controlled so that at least the first fuel injection is performed during the intake stroke, and when the occurrence of the abnormal combustion is determined by the determination means, the first fuel injection is performed. The fuel injection means is controlled so that both the fuel injection and the second fuel injection are performed during the compression stroke (Claim 1).

According to this configuration, when abnormal combustion does not occur (when it is determined that no abnormal combustion occurs), at least the first fuel injection is performed during the intake stroke, so that fuel and air are mixed before compression top dead center. can be sufficiently mixed so that the fuel can properly react with the air to properly combust these mixtures. Then, when abnormal combustion occurs (when the occurrence is determined) in which the progress of the combustion of the air-fuel mixture is rapid and the air-fuel mixture self-ignites earlier than the desired timing, the first fuel injection and the first fuel injection are performed. By performing both fuel injections in 2 during the compression stroke, the mixing time of fuel and air up to the vicinity of top dead center of compression can be shortened, and these reactions can be suppressed. Therefore, it is possible to prevent the air-fuel mixture from excessively early self-ignition near the compression top dead center and from the rapid combustion of the air-fuel mixture, even when the increase in combustion noise cannot be suppressed by changing the ignition timing. An increase in combustion noise can be suppressed.

The above configuration preferably includes an in-cylinder pressure sensor for detecting an in-cylinder pressure, and ignition means for igniting an air-fuel mixture in the cylinder at a predetermined ignition timing. The in-cylinder pressure intensity, which is the value of the spectrum of the in-cylinder pressure waveform, is equal to or greater than a predetermined determination intensity, and the amount of fuel in a predetermined proportion of the fuel supplied to the cylinder during one combustion cycle from the ignition timing. It is determined that the abnormal combustion has occurred when the index period, which is the period until the combustion is completed, is equal to or longer than a predetermined determination period.

According to this configuration, it is possible to accurately determine the degree of progress of the combustion of the air-fuel mixture and the presence or absence of abnormal combustion based on the spectrum of the pressure in the cylinder and the index period.

In the above configuration, preferably, the injection control means controls the fuel injection means so that the first fuel injection is started after the intake valve is closed when the determination means determines that the abnormal combustion has occurred. (Claim 3).

According to this configuration, the fuel injected into the combustion chamber by the first fuel injection can be prevented from diffusing throughout the combustion chamber due to the flow of intake air, and the fuel can be retained on the outer circumference of the combustion chamber. Therefore, it is possible to avoid a situation in which a large amount of highly reacted air-fuel mixture exists near the center of the combustion chamber near the compression top dead center. From this, the air-fuel mixture self-ignites excessively early near the center of the combustion chamber, where the air-fuel mixture tends to self-ignite first at a high temperature, and the high temperature facilitates the reaction of the air-fuel mixture. It is possible to prevent the air-fuel mixture from rapidly burning near the center of the combustion chamber. Therefore, an increase in combustion noise can be prevented more reliably.

In the above configuration, preferably, when the occurrence of abnormal combustion is determined by the determination means, the injection control means causes more fuel to be injected into the cylinder in one combustion cycle when the engine speed is high than when the engine speed is low. The fuel injection means is controlled so that the ratio of the amount of fuel injected into the cylinder by the second fuel injection with respect to the total amount of fuel injected is small (claim 4).

When the engine speed is high, the time per crank angle is shortened, so the mixing time of the fuel and air supplied to the cylinder by the second fuel injection is shortened. If these mixing times are too short, the fuel may not burn properly and soot may increase. On the other hand, in this configuration, when the engine speed is high, the ratio of the amount of fuel supplied to the cylinder by the second fuel injection is smaller than when the engine speed is low, so there is insufficient mixing of fuel and air. can be suppressed to prevent the increase of soot.

As an engine to which the above configuration is applied, the injection control means causes a part of the air-fuel mixture in the cylinder to undergo spark ignition combustion by ignition from an ignition means that ignites the air-fuel mixture in the cylinder, and the remaining part of the air-fuel mixture in the cylinder. There is a configuration in which fuel is injected so that the air-fuel mixture is combusted by self-ignition (Claim 5).

As described above, according to the engine control device of the present invention, it is possible to reliably prevent an increase in combustion noise.

1 is a system diagram schematically showing the overall configuration of an engine according to one embodiment of the present invention; FIG. 3 is a block diagram showing a control system of the engine; FIG. FIG. 4 is a map diagram in which engine operating regions are classified according to different operating modes; 4 is a graph showing waveforms of heat release rate and heat release amount during SPCCI combustion. FIG. 10 is a diagram showing a fuel injection pattern and heat release rate waveforms during normal control in the second operating region; (a) is a graph showing the relationship between sound frequency and spectrum, and (b) is a graph showing the relationship between cylinder pressure frequency and spectrum. FIG. 4 is a diagram schematically showing the distribution of the spectrum of the in-cylinder pressure at 7 kHz generated in the combustion chamber by cavity resonance; . FIG. 4 is a block diagram showing the flow of abnormal combustion determination; FIG. 10 is a diagram showing an example of an attenuation filter; FIG. 4 is a diagram showing changes in heat release rate with respect to crank angle; 4 is a map for explaining the flow of abnormal combustion determination; 4 is a flowchart showing the flow of abnormal combustion determination; FIG. 4 is a diagram showing a waveform of a fuel injection pattern and a heat release rate at the time of pre-ignition determination; FIG. 5 is a diagram showing the relationship between the start timings of the pre-stage injection and the post-stage injection and the engine speed in comparison between the time of normal control and the time of pre-ignition determination. FIG. 4 is a graph showing the relationship between the ratio of the injection amount of post-injection and the engine speed in comparison between normal control and pre-ignition determination; FIG. 4 is a diagram showing the relationship between the temperature of the air-fuel mixture, the equivalence ratio of the air-fuel mixture, and the susceptibility of self-ignition of the air-fuel mixture. It is a figure which showed the mode of fuel spray at the time of preignition occurrence, (a) is a figure immediately after implementation of pre-injection, (b) is a figure near compression top dead center. FIG. 2 is a diagram showing the distribution of air-fuel mixture at each temperature and each air-fuel ratio present in the combustion chamber 6, (a) is a diagram when the ratio of post-injection is 5%, and (b) is a graph when the ratio of post-injection is 5%. It is a figure at the time of 30%. 4 is a graph showing the relationship between the maximum in-cylinder pressure intensity and the reduction amount of the throttle opening during execution of intake air amount reduction control. 4 is a flowchart showing the flow of abnormal combustion avoidance control;

(1) Overall Configuration of Engine FIG. 1 is a system diagram schematically showing the overall configuration of an engine to which the engine control method and control apparatus of the present invention are applied. The engine system shown in this figure is mounted on a vehicle and includes an engine body 1 that serves as a power source for running. In this embodiment, a 4-cycle gasoline direct injection engine is used as the engine body 1 . The engine body 1 is a compression ignition engine capable of burning at least part of a mixture of fuel and air by self-ignition, as will be described later. In addition to the engine main body 1, the engine system includes an intake passage 30 through which intake air introduced into the engine main body 1 flows, an exhaust passage 40 through which exhaust gas discharged from the engine main body 1 flows, and an exhaust gas passing through the exhaust passage 40. is provided with an EGR device 50 that recirculates part of the

The engine body 1 includes a cylinder block 3 in which a cylinder 2 is formed, a cylinder head 4 attached to the upper surface of the cylinder block 3 so as to block the cylinder 2 from above, and a cylinder 2 so as to be reciprocally slidable. and a piston 5 inserted. The engine body 1 is of a multi-cylinder type having a plurality of cylinders 2 (for example, four cylinders 2 arranged in a direction perpendicular to the plane of FIG. 1). The description will focus on .

A combustion chamber 6 is defined above the piston 5, and fuel containing gasoline as a main component is supplied to the combustion chamber 6 by injection from an injector 15, which will be described later. The supplied fuel burns while being mixed with air in the combustion chamber 6, and the piston 5 pushed down by the expansion force caused by the combustion reciprocates vertically. The fuel injected into the combustion chamber 6 contains gasoline as its main component. This fuel may contain auxiliary components such as bioethanol in addition to gasoline. The injector 15 is configured to be able to inject fuel a plurality of times during one combustion cycle. In this embodiment, the injector 15 corresponds to the "fuel injection means" in the claims.

A crankshaft 7 that is an output shaft of the engine body 1 is provided below the piston 5 . The crankshaft 7 is connected to the piston 5 via a connecting rod 8 and is rotationally driven around the central axis according to the reciprocating motion (vertical motion) of the piston 5 .

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

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

The cylinder head 4 is provided with an intake port 9 and an exhaust port 10 that open to the combustion chamber 6 , an intake valve 11 that opens and closes the intake port 9 , and an exhaust valve 12 that opens and closes the exhaust port 10 . The valve format of the engine of this embodiment is a four-valve format consisting of two intake valves and two exhaust valves, and each cylinder 2 has two intake ports 9, exhaust ports 10, intake valves 11, and exhaust valves 12. are provided one by one. In this embodiment, one of the two intake ports 9 connected to one cylinder 2 is provided with a swirl valve 18 that can be opened and closed. swirling flow) is changed.

The intake valve 11 and the exhaust valve 12 are driven to open and close in conjunction with the rotation of the crankshaft 7 by valve mechanisms 13 and 14 including a pair of camshafts disposed in the cylinder head 4 .

The valve mechanism 13 for the intake valve 11 incorporates an intake VVT 13 a capable of changing at least the opening timing of the intake valve 11 . Similarly, the valve mechanism 14 for the exhaust valve 12 incorporates an exhaust VVT 14 a capable of changing at least the closing timing of the exhaust valve 12 .

In this embodiment, by controlling the intake VVT 13a and the exhaust VVT 14a, the closing timing of the exhaust valve 12 is retarded with respect to the opening timing of the intake valve 11, and both the intake valve 11 and the exhaust valve 12 are at a predetermined level. A valve overlap that opens for a period of time is realized. Also, by controlling the intake VVT 13a and the exhaust VVT 14a, the valve overlap period during which both the intake valve 11 and the exhaust valve 12 are open is changed. When the intake valve 11 and the exhaust valve 12 are driven so as to overlap each other, the burned gas is discharged from the combustion chamber 6 into at least one of the intake passage 30 and the exhaust passage 40, and then the burned gas is discharged again. Internal EGR introduced into the combustion chamber 6 is performed. As a result, burned gas (internal EGR gas) remains in the combustion chamber 6 . The amount of internal EGR gas, which is burnt gas remaining in the combustion chamber 6, varies depending on the valve overlap period, and the amount of internal EGR gas is adjusted by adjusting the valve overlap period. The intake VVT 13a (exhaust VVT 14a) may be a variable mechanism that changes only the closing timing (opening timing) while fixing the opening timing (closing timing) of the intake valve 11 (exhaust valve 12). A phase-type variable mechanism that simultaneously changes the opening timing and the closing timing of the intake valve 11 (exhaust valve 12) may be used.

In the cylinder head 4, an injector 15 for injecting fuel (mainly gasoline) into the combustion chamber 6 and a mixture of the fuel injected from the injector 15 into the combustion chamber 6 and the air introduced into the combustion chamber 6 is ignited. A spark plug 16 is provided. The cylinder head 4 is further provided with an in-cylinder pressure sensor SN3 that detects the in-cylinder pressure, which is the pressure in the combustion chamber 6 . The ignition plug 16 corresponds to the "ignition means" in the claims.

The injector 15 is a multi-hole type injector having a plurality of nozzle holes at its tip, and can radially inject fuel from the plurality of nozzle holes. The injector 15 is provided such that its tip faces the center of the crown surface of the piston 5 . In this embodiment, a cavity is formed in the crown surface of the piston 5 by recessing a region including the central portion thereof toward the opposite side (downward) of the cylinder head 4 .

The spark plug 16 is arranged at a position slightly shifted toward the intake side with respect to the injector 15 .

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

In the intake passage 30, from the upstream side thereof, an air cleaner 31 for removing foreign matter contained in the intake air (air) introduced into the combustion chamber 6 (cylinder 2), a throttle valve 32, and a filter for compressing and sending the intake air. A feeder 33, an intercooler 35 for cooling intake air compressed by the supercharger 33, and a surge tank 36 are provided. The throttle valve 32 is an opening/closing valve that opens and closes the intake passage 30, and the amount of air introduced into the combustion chamber 6 (cylinder 2) is changed according to the opening degree of the throttle valve 32.

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

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

Between the supercharger 33 and the engine body 1, an electromagnetic clutch 34 is interposed which can be electrically switched between engagement and disengagement. When the electromagnetic clutch 34 is engaged, driving force is transmitted from the engine body 1 to the supercharger 33, and supercharging by the supercharger 33 is performed. On the other hand, when the electromagnetic clutch 34 is released, the transmission of the driving force is interrupted and supercharging by the supercharger 33 is stopped.

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

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

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

The EGR device 50 has an EGR passage 51 connecting the exhaust passage 40 and the intake passage 30 , and an EGR cooler 52 and an EGR valve 53 provided in the EGR passage 51 . The EGR passage 51 connects a portion of the exhaust passage 40 downstream of the catalytic converter 41 and a portion of the intake passage 30 between the throttle valve 32 and the supercharger 33 . The EGR cooler 52 cools the exhaust gas (external EGR gas) recirculated from the exhaust passage 40 to the intake passage 30 through the EGR passage 51 by heat exchange. The EGR valve 53 is provided in the EGR passage 51 on the downstream side (closer to the intake passage 30 ) than 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 SN6 for detecting the difference between the pressure on the upstream side and the pressure on the downstream side of the EGR valve 53 .

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

Detection signals from various sensors are input to the ECU 100 . For example, the ECU 100 is electrically connected to the aforementioned crank angle sensor SN1, engine water temperature sensor SN2, cylinder internal pressure sensor SN3, air flow sensor SN4, intake air temperature sensor SN5, and differential pressure sensor SN6, and is detected by these sensors. information (that is, crank angle, engine speed, engine water temperature, cylinder internal pressure, intake air amount, intake air temperature, differential pressure across the EGR valve 53) is sequentially input to the ECU 100.

The vehicle is also provided with an accelerator sensor SN8 that detects the opening of an accelerator pedal operated by a driver who drives the vehicle.

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

The ECU 100 functionally includes an abnormal combustion determination unit 110 that determines the combustion state of the air-fuel mixture, and a fuel injection control unit 120 that controls the injector 15, as will be described later. The fuel injection control unit 120 corresponds to the "determination means", and the fuel injection control unit 120 is referred to as the "injection control means" in the claims.

(3) Basic Control FIG. 3 is a map diagram for explaining differences in operation modes according to engine speed and engine load. As shown in the figure, the operating range of the engine is broadly divided into three operating ranges: a first operating range A, a second operating range B, and a third operating range C.

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

In the first operating region A and the second operating region B, compression ignition combustion in which SI combustion and CI combustion are mixed (hereinafter referred to as SPCCI combustion) is performed. "SPCCI" in SPCCI combustion is an abbreviation for "Spark Controlled Compression Ignition."

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

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

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

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

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

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

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

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

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

In the first operating region A, the opening degree of the throttle valve 32 is fully open or nearly fully open.

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

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

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

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

(b) Second operating region In a region where the engine load is high, the amount of fuel supplied to the combustion chamber 6 is large, making it difficult to increase (lean) the air-fuel ratio of the air-fuel mixture. Accordingly, in the second operating region B, in which the engine load is higher than in the first operating region A, the air-fuel ratio in the combustion chamber 6 is made equal to or lower than the stoichiometric air-fuel ratio, and the air-fuel mixture is SPCCI-burned. In this embodiment, in the second operating region B, the air-fuel ratio of the air-fuel mixture is substantially the stoichiometric air-fuel ratio.

In the second operating region B, the opening of the throttle valve 32 is basically set so that an amount of air corresponding to the engine load is introduced into the combustion chamber 6, except when abnormal combustion avoidance control, which will be described later, is performed. be.

In the second operating region B, the injector 15 injects fuel into the combustion chamber 6 in such an amount that the air-fuel ratio becomes the stoichiometric air-fuel ratio as described above. In the present embodiment, basically, as in the first operating region A, the injector 15 controls most of the fuel to be injected into the combustion chamber 6 during one combustion cycle, except when abnormal combustion avoidance control, which will be described later, is performed. is injected during the intake stroke, and the remaining fuel is injected from the latter half of the intake stroke to the first half of the compression stroke. For example, in the second operating region B, as shown in FIG. 5, the first fuel injection (Q1) is started near the center of the compression stroke (300° CA before compression top dead center, etc.), and the intake bottom dead center BDC At (180° CA before compression top dead center), the second fuel injection (Q2) is started. The injection amount of the second fuel injection (mass of fuel injected from the injector 15 for the second time) is smaller than the injection amount of the first fuel injection (mass of fuel injected from the injector 15 for the first time). For example, the ratio of the injection amount of the second fuel injection to the total amount of fuel injected into the combustion chamber 6 during one combustion cycle is about 10%. Hereinafter, when the injector 15 injects the fuel in two portions as described above, the first fuel injection is referred to as pre-injection, and the second fuel injection is referred to as post-injection. Further, the pre-stage injection corresponds to the "first fuel injection" in the claims, and the post-stage injection corresponds to the "second fuel injection" in the claims.

Also in the second operating region B, the spark plug 16 ignites the air-fuel mixture near the compression top dead center. Also in the second operating region B, this ignition triggers the start of SPCCI combustion, a part of the mixture in the combustion chamber 6 burns due to flame propagation (SI combustion), and then the remaining mixture self-ignites. Combustion (CI combustion).

In the second operating region B, the EGR valve 53 is opened to introduce external EGR gas into the combustion chamber 6 in order to reduce NOx generated in the combustion chamber 6 . However, since a large amount of air must be introduced into the combustion chamber 6 when the engine load is high, it is necessary to reduce the amount of external EGR gas introduced into the combustion chamber 6 . Thus, in the second operating region B, the opening of the EGR valve 53 is controlled so that the amount of external EGR gas introduced into the combustion chamber 6 decreases as the load increases. Valve 53 is fully closed.

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

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

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

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

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

(4) Measures against Abnormal Combustion In the second operating region B, SPCCI combustion is performed and the engine load is high. There is a risk of sudden increases beyond the range. A sudden increase in the in-cylinder pressure results in an increase in combustion noise. Therefore, in the present embodiment, it is determined whether or not combustion has occurred in the second operating region B so that the combustion noise exceeds a desired level. Then, when such combustion occurs, control is performed to avoid continuous occurrence of this.

(4-1) Determination of Abnormal Combustion The magnitude of the cylinder internal pressure can be detected by the cylinder internal pressure sensor SN3. However, the inventors of the present application have found that in a configuration in which it is determined that abnormal combustion has occurred simply when the maximum value detected by the in-cylinder pressure sensor SN3 is higher than a predetermined value, cavity resonance that occurs in the cylinder during combustion of the air-fuel mixture It was found that the judgment accuracy was not sufficiently ensured due to

FIG. 6(a) shows the results of frequency analysis of sounds collected by a microphone around the engine body 1 when the engine body 1 is operated under first and second conditions, which are different operating conditions of the engine. is. FIG. 6(b) shows the results of frequency analysis of the in-cylinder pressure detected by the in-cylinder pressure sensor at the same time as in FIG. 6(a). In FIGS. 6A and 6B, the horizontal axis is the frequency, the vertical axis is the spectrum, the solid line is the result under the first condition, and the dashed line is the result under the second condition.

As shown in FIG. 6A, the sound spectrum under the first condition is larger than the sound spectrum under the second condition, and the noise is louder under the first condition. On the other hand, as shown in FIG. 6(b), in the region where the frequency is lower than 4 kHz, the spectrum of the in-cylinder pressure is clearly larger under the first condition than under the second condition. In this region, no difference is seen in the spectrum of the in-cylinder pressure between the first condition and the second condition except for some frequencies. Thus, the spectrum of the low-frequency component of the in-cylinder pressure, that is, the intensity/magnitude of the in-cylinder pressure, has a high correlation with noise, but the spectrum of the high-frequency component, that is, the intensity/magnitude of the in-cylinder pressure, and the noise has a small correlation. .

A high-frequency component of the waveform of the in-cylinder pressure is a wave generated by cavity resonance within the combustion chamber 6 . Therefore, in the pressure wave of the in-cylinder pressure of the high frequency component, a node region and an antinode region are formed in the combustion chamber 6 .

7(a) and 7(b) are schematic cross-sectional views of the combustion chamber 6, and diagrams schematically showing the distribution of the spectrum of the in-cylinder pressure at 7 kHz generated by cavity resonance. FIGS. 7(a) and 7(b) show distributions at predetermined timings during combustion but different from each other. In the examples of FIGS. 7(a) and 7(b), the central portion G1 in the left and right of these figures becomes a node of the pressure wave of 7 kHz, and the right end portion G3 and the left end portion G3 of FIGS. Each portion G2 becomes an antinode of the pressure wave of 7 kHz. As a result, at a predetermined timing as shown in FIG. 7(a), the spectrum at the left end G2 of the combustion chamber 6 is maximized and the spectrum at the right end G3 is minimized, but as shown in FIG. 7(b) At a timing shifted from this predetermined timing, the spectrum at the left end G2 of the combustion chamber 6 is minimized and the spectrum at the right end G3 of the combustion chamber 6 is maximized. Thus, the obtained high-frequency component in-cylinder pressure spectrum varies greatly depending on the position of the combustion chamber 6. and varies greatly depending on the installation position of the in-cylinder pressure sensor. Moreover, even with the same frequency component, the positions of the nodes and antinodes of the pressure wave change depending on the engine speed and the like.

In this way, as described above, the spectrum of the in-cylinder pressure of the high-frequency component, that is, the phenomenon that the correlation between the intensity/magnitude of the in-cylinder pressure and the noise is not ensured occurs, and abnormal combustion that causes excessive noise occurs. A configuration that simply determines whether or not there is a problem based on the maximum value of the in-cylinder pressure results in low determination accuracy.

On the other hand, the pressure wave of the in-cylinder pressure of the low-frequency component is hardly affected by the cavity resonance, so the average magnitude of the in-cylinder pressure can be detected regardless of the installation position of the in-cylinder pressure sensor. As a result, the correlation between the spectrum of the in-cylinder pressure of the low-frequency component and the noise becomes high as described above.

Based on the above findings, in the present embodiment, only the low frequency components of the in-cylinder pressure detected by the in-cylinder pressure sensor SN3 are used to determine whether abnormal combustion has occurred.

Abnormal combustion determination section 110 determines whether or not abnormal combustion has occurred. FIG. 8 is a block diagram showing the flow of abnormal combustion determination performed by abnormal combustion determination section 110. As shown in FIG. Abnormal combustion determination section 110 functionally includes bandpass filter section 111, spectrum calculation section 112, attenuation filter section 113, maximum in-cylinder pressure strength calculation section 114, comparison determination section 115, and combustion center-of-gravity timing calculation section. 116 and an index period calculator 117 .

The bandpass filter unit 111 extracts only low frequency components from the in-cylinder pressure detected by the in-cylinder pressure sensor SN3. In the present embodiment, only the frequency components below the first frequency and above the second frequency are extracted from the in-cylinder pressure detected by the in-cylinder pressure sensor SN3. The first frequency is set near the maximum frequency of the in-cylinder pressure that is not affected by cavity resonance. The second frequency is set to a value near the maximum frequency of the waveform of the in-cylinder pressure that is not caused by combustion, that is, the in-cylinder pressure that occurs with the reciprocating motion of the piston 5 . For example, the first frequency is set to 4 kHz and the second frequency is set to 1 kHz. It should be noted that sounds with a frequency higher than 4 kHz are relatively difficult for human ears to perceive. Therefore, when the first frequency is set to 4 kHz, it is possible to ensure the correlation between the noise perceived by humans and the level of the in-cylinder pressure.

Spectrum calculation section 112 performs Fourier analysis on the low-frequency component of the in-cylinder pressure extracted by band-pass filter section 111 to calculate the spectrum of each frequency (in the range from the second frequency to the first frequency). Before the Fourier analysis is performed by the spectrum calculating section 112, processing is performed to remove pressure changes, that is, ignition noise caused by the driving of the spark plug 16 from the low-frequency components of the in-cylinder pressure extracted by the band-pass filter section 111. You may do so.

Attenuation filter section 113 applies an attenuation filter to the spectrum calculated by spectrum calculation section 112 . That is, part of the pressure wave generated in the combustion chamber 6 is attenuated by the cylinder block 3 and the like, and the level of sound transmitted as noise to the outside of the engine body 1 is reduced. Attenuation filter section 113 applies an attenuation filter that simulates such attenuation to the spectrum calculated by spectrum calculation section 112 . In this embodiment, the attenuation filter section 113 subtracts the attenuation amount set for each frequency as shown in FIG. 9 from the spectrum of each frequency calculated by the spectrum calculation section 112 . Specifically, the attenuation filter in FIG. 9 is set so that the attenuation amount decreases as the frequency increases toward the first frequency. The higher the value, the smaller the value will be subtracted.

Maximum in-cylinder pressure intensity calculation section 114 calculates the maximum value of the spectrum of the in-cylinder pressure waveform from the second frequency to the first frequency (the second frequency or more and the first frequency or less) after being processed by attenuation filter section 113. It is calculated as the in-cylinder pressure strength CPLF. In this embodiment, the spectrum of the in-cylinder pressure after being processed by the attenuation filter unit 113 is subjected to 1/3 octave band processing, the spectrum of each band is calculated, and then the maximum value of the spectrum of all bands is extracted. . The 1/3 octave banding process is a process of dividing each octave region of the frequency spectrum (the region from a given frequency to its double frequency) into three, and calculating the in-cylinder pressure level of each divided band. . For example, the 1/3 octave banding process identifies the spectrum of each band with center frequencies of 1 kHz, 1.25 kHz, 1.6 kHz, 2 kHz, 2.5 kHz, 3.15 kHz, and 4 kHz, respectively. Thus, in the present embodiment, the maximum value of the spectrum of the in-cylinder pressure of the low-frequency component after being attenuated by the cylinder block 3 or the like is calculated. The maximum in-cylinder pressure intensity CPLF corresponds to the "in-cylinder pressure intensity" in the claims.

Comparison determination unit 115 compares maximum in-cylinder pressure intensity CPLF calculated by maximum in-cylinder pressure intensity calculation unit 114 with a preset determination intensity. Then, when the maximum in-cylinder pressure intensity CPLF exceeds the determination intensity, comparison determination unit 115 determines that combustion has occurred in which the combustion noise exceeds the desired level. It is known that the maximum in-cylinder pressure intensity CPLF is 80 dB or more when such combustion with combustion noise exceeding a desired level occurs. From this, the determination strength is set to 80 dB. In other words, in the present embodiment, combustion with a maximum in-cylinder pressure strength CPLF of less than 80 dB is defined as normal combustion, and combustion with a maximum in-cylinder pressure strength CPLF of 80 dB or more is combustion noise exceeding a desired level. defined as being

FIG. 10 compares the heat release rate (solid line) when normal combustion is realized and the heat release rate (broken line, dashed line) when combustion occurs where the maximum in-cylinder pressure intensity CPLF is 80 dB or more. showing. As shown in FIG. 10, when combustion occurs with a maximum in-cylinder pressure intensity CPLF of 80 dB or more, combustion starts earlier and progresses faster than normal combustion. However, even when the maximum in-cylinder pressure strength CPLF is 80 dB or more, the timing at which the heat release rate rises sharply, that is, the timing at which CI combustion starts is not constant, as indicated by the broken and chain lines in FIG. .

In the case of combustion in which CI combustion starts relatively late, as indicated by the dashed line, the progress of combustion of the air-fuel mixture can be retarded by changing the ignition timing. Specifically, the combustion speed can be slowed by retarding the ignition timing to cause main combustion at a later timing in the expansion stroke. On the other hand, in the case where CI combustion starts relatively early and the progress of combustion is excessively fast, as indicated by the dashed line, even if the ignition timing is changed, the progress of combustion of the air-fuel mixture cannot be changed. In other words, in the combustion shown by the dashed line, the reaction between the fuel and air has sufficiently progressed before the ignition timing, so even if the ignition timing is changed, the degree of combustion of the air-fuel mixture cannot be changed. The air-fuel mixture self-ignites early regardless of the ignition timing.

Correspondingly, when abnormal combustion determining section 110 determines that combustion with maximum in-cylinder pressure intensity CPLF equal to or greater than the determination intensity has occurred, it further determines which pattern of combustion this combustion is. do. That is, comparison determination unit 115 determines whether the combustion when maximum in-cylinder pressure strength CPLF is equal to or greater than the determination strength is the first abnormal combustion in which the progress of the combustion of the air-fuel mixture can be changed by changing the ignition timing. It is determined whether it is the second abnormal combustion in which the progress of the combustion of the air-fuel mixture cannot be changed by changing the . Hereinafter, the first abnormal combustion will be referred to as knock, and the second abnormal combustion will be referred to as pre-ignition. In this embodiment, this second abnormal combustion, that is, pre-ignition, corresponds to "abnormal combustion" in the claims.

In the present embodiment, as shown in FIG. 11, when the abnormal combustion determination unit 110 determines that combustion occurs in which the maximum in-cylinder pressure strength CPLF is equal to or greater than the determination strength, the index period from the ignition timing to the combustion center-of-gravity timing is Pre-ignition is determined when it is shorter than a predetermined determination period, and knocking is determined when the index period from ignition timing to combustion center-of-gravity timing is equal to or longer than the determination period. The combustion center-of-gravity timing is the timing at which 50% of the total amount (mass) of fuel supplied to the combustion chamber 6 during one combustion cycle completes combustion. In this embodiment, 50% of the total amount (mass) of the fuel corresponds to "a predetermined proportion of the fuel supplied to the cylinder during one combustion cycle".

Combustion center-of-gravity timing calculator 116 calculates the volume of combustion chamber 6 based on the in-cylinder pressure detected by in-cylinder pressure sensor SN3, the crank angle detected by crank angle sensor SN1, and the volume of combustion chamber 6 at each crank angle stored in ECU 100. Then, the heat release rate at each crank angle is calculated, and the heat release amount is calculated by integrating this. Then, as shown in FIG. 4, the amount of heat release at a predetermined crank angle is assumed to be 100%, and the crank angle at which 50% of the amount of heat release occurs is calculated as the combustion center-of-gravity timing.

The index period calculator 117 calculates the period from the ignition timing (the ignition timing command value issued from the ECU 100 to the spark plug 16) to the combustion center of gravity timing calculated by the combustion center of gravity timing calculator 116 as the index period. .

The comparison determination unit 115 compares the index period calculated by the index period calculation unit 117 and the determination period. When it is determined that the combustion in which the maximum in-cylinder pressure intensity CPLF is equal to or greater than the determination intensity has occurred, comparison determination unit 115 determines that knock has occurred when the index period is less than the determination period.

In addition, when the combustion cycle in which the maximum in-cylinder pressure intensity CPLF is equal to or greater than the determination intensity and the index period is less than the determination period occurs continuously in one cylinder 2 for a predetermined number of times or more, pre-ignition occurs. determined to have occurred. That is, in the present embodiment, it is not determined that pre-ignition has occurred if only one combustion cycle occurs in which the maximum in-cylinder pressure intensity CPLF is equal to or greater than the determination intensity and the index period is less than the determination period. It is determined that pre-ignition has occurred for the first time in the same cylinder 2 for the number of determination times or more. The determination period and the number of times of determination are set in advance and stored in the comparison/determination unit 115 . For example, the number of determinations is set to two. It has been found that pre-ignition occurs when the index period is less than 10° CA. Accordingly, the determination period is set to 10°CA.

Using the flow chart of FIG. 12, the procedure for judging abnormal combustion is summarized as follows. Steps S1 to S19 shown in FIG. 12 are performed for each cylinder 2 each time combustion in each cylinder 2 is completed.

In step S1, the abnormal combustion determination unit 110 reads the in-cylinder pressure and the like detected by the in-cylinder pressure sensor SN3.

Next, in step S<b>2 , the abnormal combustion determination unit 110 determines whether or not the current operating point of the engine is within the second operating region B.

If the determination in step S2 is NO and the current operating point is not within the second operating region B, the process proceeds to step S19. In step S19, abnormal combustion determination unit 110 sets counter (i) to zero. Counters will be described later.

On the other hand, if the determination in step S2 is YES and the current operating point is within the second operating region B, that is, if the engine is being operated in the second operating region B, the process proceeds to step S3. .

In step S3, abnormal combustion determination unit 110 calculates maximum in-cylinder pressure intensity CPLF as described above based on the in-cylinder pressure detected by in-cylinder pressure sensor SN3. After step S3, the process proceeds to step S4.

In step S4, abnormal combustion determination unit 110 calculates the index period, that is, the period from ignition timing to combustion center-of-gravity timing, as described above, based on the in-cylinder pressure and the crank angle. After step S4, the process proceeds to step S5.

At step S5, abnormal combustion determination unit 110 determines whether or not the maximum in-cylinder pressure strength CPLF calculated at step S3 is less than the determination strength.

When the determination in step S5 is YES and the maximum in-cylinder pressure strength CPLF is less than the determination strength, the process proceeds to step S6. In step S6, the abnormal combustion determination unit 110 sets the counter (i) to 0, proceeds to step S7, determines that normal combustion is being performed, and terminates the process (returns to step S1).

On the other hand, when the determination in step S5 is NO and the maximum in-cylinder pressure strength CPLF is greater than or equal to the determination strength, the process proceeds to step S8. In step S8, abnormal combustion determination unit 110 determines whether or not the index period calculated in step S4 is shorter than the determination period.

If the determination in step S8 is YES and the index period is less than the determination period, the process proceeds to step S9. In step S9, the abnormal combustion determination unit 110 sets the counter (i) to a value obtained by adding 1 to the counter (i-1) calculated in the previous calculation cycle. After step S9, the process proceeds to step S10.

In step S10, abnormal combustion determination unit 110 determines whether counter (i) is equal to or greater than the number of times of determination. If the determination in step S10 is NO and the counter (i) is less than the number of determinations, the abnormal combustion determination unit 110 ends the process (returns to step S1).

On the other hand, if the determination in step S10 is YES and the counter (i) is equal to or greater than the determination count, the process proceeds to step S11, and the abnormal combustion determination unit 110 determines that preignition has occurred.

Thus, the counter is a parameter representing the number of consecutive combustion cycles in which the maximum in-cylinder pressure intensity CPLF is greater than or equal to the determination intensity and the index period is less than the determination period in one cylinder. Therefore, as described above, this counter indicates that the condition is not satisfied when the engine is not operated in the second operating region B, or even when the engine is operating in the second operating region B. In that case, it will be 0. Then, as described above, in the present embodiment, when this condition continues for one cylinder 2 more than the number of determination times, it is determined that pre-ignition has occurred.

Returning to step S8, when the determination in step S8 is NO and the index period is equal to or longer than the determination period, the process proceeds to step S12. In step S12, abnormal combustion determination unit 110 sets counter (i) to zero. After step S12, the process proceeds to step S13. In step S13, abnormal combustion determination unit 110 determines that knock has occurred, and ends the process (returns to step S1).

(4-2) Abnormal Combustion Avoidance Control Next, abnormal combustion avoidance control, which is control for avoiding knocking and pre-ignition from occurring in succession, will be described.

(a) Knock Avoidance Control When the abnormal combustion determination unit 110 determines that knock has occurred, the ECU 100 retards the ignition timing. That is, the ECU 100 controls the spark plug 16 so that ignition is performed at a timing that is more retarded than when it is determined that knocking has not occurred. Specifically, when it is determined that knock has occurred, the ECU 100 retards the ignition timing of the next combustion cycle from the ignition timing one combustion cycle before by a predetermined timing. At the time of knock determination, the injection pattern of fuel injection is controlled in the same manner as during normal control and when the maximum in-cylinder pressure intensity CPLF is less than the determination intensity.

(b) Pre-ignition avoidance control When pre-ignition is determined, that is, when abnormal combustion determination unit 110 determines that pre-ignition has occurred, ECU 100 (fuel injection control unit 120) delays the execution timing of fuel injection, that is, the injection timing. Then, the fuel injection pattern is switched to a pattern different from that during normal control (when it is not determined that pre-ignition has occurred).

As described above, during normal control, the injector 15 performs pre-injection in the intake stroke (Q1), and performs post-injection (Q2) from the latter half of the intake stroke to the first half of the compression stroke. On the other hand, during the pre-ignition determination, the ECU 100 causes the injector 15 to inject the fuel in two steps as in the case of normal control. The injector 15 is driven so that the injection timing) is retarded relative to that during normal control.

As indicated by the dashed line in FIG. 13 and as described above, during normal control, the pre-injection (Q1) is started during the intake stroke in the second operating region B, and the post-injection (Q2) is started during the intake stroke. It starts from the latter half to the first half of the compression stroke. On the other hand, as shown by the solid line in FIG. 13, during the pre-ignition determination, the pre-injection (Q11) starts in the first half of the compression stroke (180° CA before top dead center of compression to 90° CA before top dead center of compression). Then, the post-injection (Q12) is started in the latter half of the compression stroke (90° CA before compression top dead center to compression top dead center). In addition, the start timing of the pre-injection (Q11) is retarded relative to the closing timing IVC of the intake valve 11 . For example, at the time of pre-ignition determination, the front injection (Q11) is started at 180°CA before the compression top dead center, and the rear injection (Q12) is started at about 10 to 20°CA before the compression top dead center. Also, the start timings of the pre-injection (Q1) and the post-injection (Q2) during normal control are set with respect to the engine speed as indicated by the dashed lines in FIG. The start timing of Q11, Q12) is set as indicated by the solid line in FIG.

Further, during the pre-ignition determination, the ECU 100 increases the ratio of the injection amount of the post-injection.

Specifically, the ratio of the injection amount of the post-stage injection to the total amount of fuel injected into the combustion chamber 6 during one combustion cycle is made larger during pre-ignition determination than during normal control. For example, while the ratio of the injection amount of the post-stage injection during normal control is about 10%, it is about 20 to 30% during pre-ignition determination. Further, as shown by the dashed line in FIG. 15, the ratio of the injection amount of the post-injection (Q2) during normal control is maintained constant regardless of the engine speed, whereas as shown by the solid line in FIG. , the ratio of the injection amount of the post-injection (Q12) at the time of pre-ignition determination becomes smaller as the engine speed increases. Hereinafter, the fuel injection pattern at the time of pre-ignition determination is referred to as a pre-ignition avoidance pattern.

When the execution timing of fuel injection is retarded, the mixing time of fuel and air and the reaction time between them until reaching the compression top dead center are shortened, and the progress of the reaction of the air-fuel mixture near the compression top dead center is reduced. Become slow. Accordingly, if the fuel injection pattern is set to the pre-ignition avoidance pattern and the implementation timings of the pre- and post-injections are retarded, the reaction proceeds before the start of ignition and the air-fuel mixture that is in a state just before self-ignition burns. A situation in which a large amount exists in the chamber 6 is avoided, and the occurrence of preignition is suppressed.

Further, in the pre-ignition avoidance pattern, pre-injection is performed in the first half of the compression stroke and after the intake valve 11 is closed, thereby more reliably preventing premature self-ignition of the air-fuel mixture. A specific description will be given with reference to FIG. 16 and the like.

FIG. 16 shows the results of examining the relationship between the temperature of the air-fuel mixture, the equivalence ratio of the air-fuel mixture (the value obtained by dividing the stoichiometric air-fuel ratio by the air-fuel ratio), and the susceptibility of self-ignition of the air-fuel mixture. The horizontal axis of the graph in FIG. 16 represents the temperature of the air-fuel mixture, and the vertical axis represents the susceptibility to self-ignition. Specifically, the vertical axis represents the amount of OH radicals required to cause the mixture to self-ignite. can be said to be difficult to self-ignite. The three lines in the graph of FIG. 16 are lines in which the equivalence ratio (air-fuel ratio) of the air-fuel mixture differs from each other.

As shown in FIG. 16, the higher the temperature in the cylinder, that is, the higher the temperature of the air-fuel mixture, the easier it is for the air-fuel mixture to self-ignite. Here, the central portion of the combustion chamber 6 becomes hotter than the outer peripheral portion. As a result, self-ignition of the air-fuel mixture begins in the central portion of the combustion chamber 6 first. In contrast, in the present embodiment, the pre-injection (Q11) and the post-injection (Q12) are performed at the timing described above, thereby excessively promoting the combustion of the air-fuel mixture in the central portion of the combustion chamber 6. to be prevented.

Specifically, the pre-injection (Q11) is performed during the first half of the compression stroke and after the intake valve 11 is closed, as the piston 5 is rising and after the intake valve 11 is closed. weak flow. Therefore, by performing the pre-injection (Q11) in the first half of the compression stroke and after the intake valve 11 is closed, the diffusion of the fuel F1 injected by the pre-injection (Q11) is It is suppressed, and most of this fuel F1 stays near the outer circumference of the combustion chamber 6 . Then, as shown in FIG. 17(b), this fuel distribution is generally maintained until near the compression top dead center. Therefore, the mixture and reaction of the fuel and air associated with the pre-injection (Q11) mainly occur near the outer circumference of the combustion chamber 6, and the amount of temperature increase near the center of the combustion chamber 6 due to this reaction is kept small. In addition, the amount of the air-fuel mixture in the state just before self-ignition distributed in the center of the combustion chamber 6 near the top dead center of the compression can be reduced. Therefore, the timing of the self-ignition of the air-fuel mixture that starts near the center of the combustion chamber 6 is delayed, the progress of combustion is delayed, and the occurrence of preignition is suppressed. Note that, during normal control, the pre-injection (Q1) is performed during the intake stroke, so that the pre-injected fuel diffuses throughout the combustion chamber 6 .

Here, in the pre-ignition avoidance pattern, post-injection is performed in the latter half of the compression stroke. In the second half of the compression stroke, when the piston 5 is injected at a timing close to the compression top dead center, the fuel is supplied to the vicinity of the center of the combustion chamber 6 . Also, in the pre-ignition avoidance pattern, the ratio of the injection amount of the post-stage injection is increased. Therefore, in the pre-ignition avoidance pattern, the air-fuel ratio of the air-fuel mixture near the center of the combustion chamber 6 becomes small (rich) near the top dead center of the compression stroke. It can be said that it becomes easy to self-ignite. However, as described above, in the pre-ignition avoidance pattern, the amount of temperature increase in the vicinity of the center of the combustion chamber 6 due to the pre-injection is kept small. In addition, as described above, due to the short time from the post-injection to the compression top dead center, the reaction between the fuel and the air related to the post-injection does not sufficiently progress until near the compression top dead center. Also, the temperature rise in the vicinity of the center of the combustion chamber 6 is suppressed. Therefore, the self-ignition of the air-fuel mixture is not excessively promoted in the vicinity of the center of the combustion chamber 6, and the progress of the combustion of the air-fuel mixture slows down.

This is also clear from a comparison between FIGS. 18(a) and 18(b). 18(a) and 18(b) are diagrams showing the distribution of air-fuel mixture at each temperature and each air-fuel ratio in the combustion chamber 6 at compression top dead center. In FIGS. 18(a) and 18(b), the darker the color, the higher the distribution ratio of the air-fuel mixture. FIG. 18(a) is a diagram when the ratio of post-injection is 5% (when the ratio of pre-injection is 95%), and FIG. 18(b) is a graph when the ratio of post-injection is 30%. FIG. 10 is a diagram when the ratio of the pre-injection is set to 70%. The dashed lines shown in FIGS. 18(a) and 18(b) are lines connecting temperatures and air-fuel ratios at which the susceptibility to self-ignition becomes the same. As shown in FIG. 18(a), when the ratio of the pre-injection is large and the ratio of the post-injection is small, the ratio of the air-fuel ratio is relatively large (lean) but the temperature is high. On the other hand, as shown in FIG. 18B, when the ratio of the pre-injection is small and the ratio of the post-injection is large, the ratio of the air-fuel ratio is relatively small (rich) but the temperature is low. becomes larger. Further, in the case of FIG. 18(b), more air-fuel mixture is distributed in the region where self-ignition is difficult to occur than in the case of FIG. 18(a). Therefore, in the pre-ignition avoidance pattern, although the ratio of the post-injection is increased and an air-fuel mixture with a high air-fuel ratio is formed near the center of the combustion chamber 6, the air-fuel mixture is in a state where self-ignition is difficult, and the combustion of the air-fuel mixture is delayed. Progress is slowed.

Further, in the present embodiment, the ECU 100 reduces the amount of intake air introduced into the combustion chamber 6 when pre-ignition is determined and when a predetermined condition, which will be described later, is satisfied. Specifically, the opening degree of the throttle valve 32 is reduced (closed). In this embodiment, when it is determined that pre-ignition has occurred in a state where a predetermined condition is satisfied, the opening of the throttle valve 32 is made smaller than the opening just before this determination is made. As shown in FIG. 19, the ECU 100 increases the amount of reduction in the opening of the throttle valve 32 as the maximum in-cylinder pressure strength CPLF increases, and decreases the opening of the throttle valve 32 as the maximum in-cylinder pressure strength CPLF increases. to increase the reduction amount of the intake air amount. During pre-ignition determination and when a predetermined condition is satisfied, the air-fuel ratio of the air-fuel mixture in the combustion chamber 6 is maintained near the stoichiometric air-fuel ratio in the same manner as during normal control, and the total amount of fuel supplied to the combustion chamber 6 along with the amount of intake air is maintained. reduced. As a result, the amount of heat generated in the combustion chamber 6 is reduced and the occurrence of preignition is suppressed.

Here, the throttle valve 32 is provided in the intake passage 30 communicating with all the cylinders 2, and when the opening degree of the throttle valve 32 is changed, the intake air amount of all the cylinders 2 is changed. Further, due to the delay in driving the throttle valve 32 and the delay in transportation of gas, the amount of air flowing through the intake passage 30 only gradually decreases even if the opening degree of the throttle valve 32 is changed. Accordingly, after it is determined that pre-ignition has occurred in a predetermined cylinder 2 and the opening degree of the throttle valve 32 is reduced, if it is determined that pre-ignition has occurred in another cylinder 2 and the opening degree of the throttle valve 32 is further reduced, Inspiratory volume may be excessively low. Therefore, in the present embodiment, further changes in the opening of the throttle valve 32 are stopped from when the opening of the throttle valve 32 is reduced until combustion is completed once in all cylinders 2 . For example, in a 4-cylinder engine, the change in the opening of the throttle valve 32 is prohibited within 720° CA after the opening of the throttle valve 32 is reduced.

Reducing the total amount of air intake and fuel reduces engine torque. Therefore, the control for reducing the intake air amount is performed only when the state in which it is difficult to suppress the occurrence of pre-ignition only by changing the fuel injection pattern continues.

Specifically, when the maximum in-cylinder pressure intensity CPLF is particularly large, it becomes difficult to suppress the occurrence of pre-ignition only by changing the fuel injection pattern.

Here, when the operating point of the engine changes and the preset target EGR rate changes, the target amount of EGR gas introduced into the combustion chamber 6 may be reduced due to delay in response of the EGR valve 53 and delay in transportation of gas. values may deviate. The EGR rate is the amount (mass) of EGR gas in the combustion chamber 6 with respect to the total amount (total mass) of gas in the combustion chamber 6 . If the deviation occurs, the amount of intake air introduced into the combustion chamber 6 instead of EGR gas increases, the air-fuel ratio of the air-fuel mixture increases (becomes lean), and preignition may occur. Alternatively, the amount of EGR gas introduced into the combustion chamber 6 may become excessively large, causing the temperature in the combustion chamber 6 to rise and preignition to occur. However, the pre-ignition that occurs in this manner is eliminated by eliminating the delay in the response of the EGR valve 53 and the delay in transportation of the gas, so there is little need to reduce the intake air amount.

Therefore, in the present embodiment, when it is determined that pre-ignition has occurred, the absolute value of the deviation of the EGR rate from the target value is equal to or less than the predetermined determination value, and the maximum in-cylinder pressure strength CPLF is the above-described determination value. The amount of intake air introduced into the combustion chamber 6 is reduced when the intensity is equal to or greater than the reference intensity, which is greater than the intensity. That is, the predetermined conditions are set such that the absolute value of the deviation of the EGR rate from the target value is equal to or less than the judgment value and the maximum in-cylinder pressure intensity CPLF is equal to or greater than the reference intensity.

The above abnormal combustion avoidance control can be summarized as shown in the flow chart of FIG.

In step S21, ECU 100 determines whether or not abnormal combustion determination unit 110 has determined that knocking has occurred. When this determination is YES and it is determined that knocking has occurred, the process proceeds to step S22. In step S22, the ECU 100 retards the ignition timing and terminates the process (returns to step S1).

On the other hand, if the determination in step S21 is NO and it is not determined that knocking has occurred, the process proceeds to step S23. In step S23, the ECU 100 determines whether or not the abnormal combustion determination unit 110 has determined that pre-ignition has occurred. When this determination is NO and it is not determined that pre-ignition has occurred, the process proceeds to step S30. In step S30, the abnormal combustion avoidance control is not performed, and the ECU 100 performs normal control.

On the other hand, if the determination in step S23 is YES and it is determined that pre-ignition has occurred, the process proceeds to step S24, and the ECU 100 increases the ratio of the injection amount of the post-stage injection, Injection implementation timing (start timing) is retarded more than during normal control.

If the determination in step S23 is YES, the ECU 100 calculates the deviation of the EGR rate from the target value in step S26.

Specifically, a target EGR rate, which is a target value of the EGR rate, is preset and stored in ECU 100 . For example, the target EGR rate is preset according to the engine speed, the engine load, etc., and is stored in the ECU 100 as a map, and the ECU 100 extracts values corresponding to the current engine speed and the engine load from this map. do. The ECU 100 also estimates the EGR rate of the air-fuel mixture based on the intake air amount detected by the airflow sensor SN4 and the differential pressure across the EGR valve 53 detected by the differential pressure sensor SN6. In step S26, the ECU 100 calculates the difference between the estimated EGR rate and the target EGR rate, that is, the deviation of the EGR rate.

After step S26, the process proceeds to step S27. In step S27, the ECU 100 determines whether or not the conditions that the absolute value of the deviation of the EGR rate calculated in step S26 is less than the determination value and the maximum in-cylinder pressure strength CPLF is greater than or equal to the reference strength are established. judge. The aforementioned determination value and reference intensity are preset and stored in the ECU 100 .

If the determination in step S27 is NO and the condition is not met, the process is terminated (returns to step S1). On the other hand, when the determination in step S27 is YES and the above conditions are satisfied, the process proceeds to step S28.

In step S28, the ECU 100 satisfies a condition that the intake air amount reduction control in step S29, which will be described later, has not yet been performed, or that combustion in all cylinders 2 has ended after the intake air amount reduction control was started. Determine whether or not If the determination in step S28 is NO, that is, the intake air amount reduction control has already been performed, and combustion in all cylinders 2 has not ended after the intake air amount reduction control was started, the process ends. (return to step S1). On the other hand, if the determination in step S28 is YES and the intake air amount reduction control has not yet been performed, or if combustion in all cylinders 2 has ended after the intake air amount reduction control was started, step Proceed to S29. In step S29, the ECU 100 performs intake air amount reduction control to reduce the intake air amount and ends the process (returns to step S1). Specifically, the ECU 100 reduces the amount of intake air introduced into each combustion chamber 6 by reducing the opening degree of the throttle valve 32 (closed side), as described above.

(5) Operation, Etc. As described above, in the present embodiment, during normal control and during knock determination, when the maximum in-cylinder pressure strength CPLF is less than the determination strength or when the maximum in-cylinder pressure strength CPLF is equal to or greater than the determination strength, When the index period is shorter than the determination period (that is, when pre-ignition does not occur), the pre-injection (Q1) is performed during the intake stroke. Therefore, the fuel and air can be sufficiently mixed until near the top dead center of the compression stroke, and the fuel can appropriately react with the air. As a result, fuel consumption performance and exhaust performance can be improved.

When preignition occurs, that is, when the maximum in-cylinder pressure intensity CPLF is equal to or greater than the determination intensity and the index period is shorter than the determination period, the progress of the combustion of the air-fuel mixture is particularly rapid, and the ignition timing is retarded. Both the pre-injection (Q11) and the post-injection (Q12) are performed during the compression stroke when abnormal combustion that cannot be avoided even by the compression stroke occurs. Therefore, in this case, the mixing time and the reaction time of the fuel and air until near the top dead center of the compression can be shortened, and these reactions can be suppressed until near the top dead center of the compression. Therefore, it is possible to prevent the air-fuel mixture from excessively early self-ignition and rapid combustion of the air-fuel mixture near the compression top dead center. Further, it is possible to prevent combustion noise from increasing.

In particular, in this embodiment, when it is determined that pre-ignition has occurred, the pre-injection (Q11) is performed after the intake valve 11 is closed. Therefore, as described above, the fuel related to the pre-injection (Q11) can be prevented from diffusing throughout the combustion chamber 6, and the fuel can be retained on the outer periphery of the combustion chamber 6. Therefore, it is possible to avoid a situation in which a large amount of a highly reacted air-fuel mixture exists near the center of the combustion chamber 6 near the top dead center of the compression stroke. Prevents rapid combustion.

Further, when the engine speed is high, the time per crank angle is shortened, so the mixing time of the fuel and air supplied to the cylinder by the post-injection (Q12) is shortened. If these mixing times are too short, the fuel may not burn properly and soot may increase. In contrast, in the present embodiment, when it is determined that pre-ignition has occurred, the injection amount of the post-injection (Q12) supplied to the cylinder in one combustion cycle is greater when the engine speed is high than when the engine speed is low. is reduced. Therefore, insufficient mixing of fuel and air can be suppressed, and an increase in soot can be prevented.

Further, in this embodiment, it is determined whether or not preignition has occurred based on the spectrum of the in-cylinder pressure detected by the in-cylinder pressure sensor SN3 and the index period from the ignition timing to the combustion center of gravity timing. Therefore, it is possible to accurately determine whether or not preignition, which is abnormal combustion in which the degree of progress of combustion is excessively high, has occurred. By changing the fuel injection pattern based on this determination result, this can be changed appropriately.

(6) Modification In the above embodiment, when a combustion cycle in which the maximum in-cylinder pressure intensity CPLF is equal to or greater than the determination intensity and the index period is equal to or greater than the determination period in one cylinder continues a plurality of times (determined number of times or more), Although the case where it is determined that pre-ignition has occurred has been described, it may be determined that pre-ignition has occurred when the combustion cycle has occurred only once. However, with the configuration of the above embodiment, it is possible to avoid determination that pre-ignition has occurred due to an erroneous calculation of the maximum in-cylinder pressure intensity CPLF and the index period. Also, when the maximum in-cylinder pressure strength CPLF temporarily exceeds the determination strength and the index period exceeds the determination period, it is determined that pre-ignition has occurred, and thereafter the fuel injection pattern is changed. can be avoided, and the chances of changing the fuel injection pattern can be reduced.

In the above-described embodiment, the post-injection (Q2) is performed during the period from the latter half of the intake stroke to the first half of the compression stroke during normal control and knock control. As in (Q1), it may be performed during the intake stroke.

In the above embodiment, the case where the engine performs SPCCI combustion has been described, but the combustion mode performed in the engine is not limited to this. Further, the above-described abnormal combustion determination and abnormal combustion avoidance control may be performed outside the second operating region B.

In the above embodiment, the maximum value of the spectrum of the in-cylinder pressure from the second frequency to the first frequency (maximum in-cylinder pressure Intensity CPLF) has been described. Instead of this, the average value of the spectral values of the in-cylinder pressure from the second frequency to the first frequency may be used as the parameter. Further, whether or not abnormal combustion has occurred may be determined based on the spectrum of the in-cylinder pressure, which is the frequency component of the first frequency or lower and also includes the frequency component of the second frequency or lower.

1 engine body 2 cylinder 6 combustion chamber 15 injector (fuel supply means)
16 spark plug (ignition means)
100 ECUs
110 Abnormal Combustion Determination Unit (Determination Means)
120 fuel injection control unit (injection control means)
SN6 Cylinder pressure sensor

Claims (5)

Controls a compression ignition engine that includes fuel injection means for injecting fuel into a cylinder and can burn at least part of a mixture of fuel and air injected from the fuel injection means by self-ignition. A device for
Determination means for determining the presence or absence of abnormal combustion in which the progress of combustion of the air-fuel mixture is excessively fast;
Injection control means for causing the fuel injection means to perform a first fuel injection and a second fuel injection whose injection timing is later than the first fuel injection,
The injection control means is
controlling the fuel injection means so that at least the first fuel injection is performed during an intake stroke when the determination means determines that the abnormal combustion does not occur;
controlling the fuel injection means so that both the first fuel injection and the second fuel injection are performed during a compression stroke when the determination means determines that the abnormal combustion has occurred; An engine control device characterized by: In the engine control device according to claim 1,
an in-cylinder pressure sensor that detects the in-cylinder pressure, which is the pressure in the cylinder;
Equipped with ignition means for igniting the air-fuel mixture in the cylinder at a predetermined ignition timing,
The determination means determines that an in-cylinder pressure intensity, which is a spectrum value of a waveform of the in-cylinder pressure detected by the in-cylinder pressure sensor, is equal to or greater than a predetermined determination intensity, and that the in-cylinder pressure is applied to the cylinder during one combustion cycle from the ignition timing. It is determined that the abnormal combustion has occurred when an index period, which is a period until combustion of a predetermined proportion of the supplied fuel is completed, is equal to or greater than a predetermined determination period. engine controller. The engine control device according to claim 1 or 2,
The injection control means controls the fuel injection means so that the first fuel injection is started after the intake valve is closed when the determination means determines that the abnormal combustion has occurred. engine controller. In the engine control device according to any one of claims 1 to 3,
The injection control means controls the total amount of fuel injected into the cylinder in one combustion cycle when the occurrence of abnormal combustion is determined by the determination means to be greater when the engine speed is high than when the engine speed is low. 2. An engine control device, wherein the fuel injection means is controlled so that the proportion of the amount of fuel injected into the cylinder by the fuel injection of 2 is reduced. In the engine control device according to any one of claims 1 to 4,
The injection control means causes a portion of the air-fuel mixture in the cylinder to undergo spark ignition combustion by ignition from an ignition means that ignites the air-fuel mixture in the cylinder, and the remaining air-fuel mixture burns by self-ignition. An engine control device characterized by injecting fuel.

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