JP7238571B2 - Engine control method and engine control device - Google Patents

Description

The present invention comprises an engine body in which a cylinder is formed, and an ignition device for igniting an air-fuel mixture in the cylinder. The present invention relates to a control method and a control device for a compression ignition engine capable of executing partial compression ignition combustion in which the remaining air-fuel mixture is burned by self-ignition.

2. Description of the Related Art Engines installed in vehicles and the like are required to have improved fuel efficiency. In order to improve the fuel efficiency, it is desirable to burn the air-fuel mixture at an appropriate time near the compression top dead center. On the other hand, for example, Patent Document 1 discloses a spark ignition type engine in which an air-fuel mixture is ignited and burned by flame propagation, and the ignition timing is adjusted so that the combustion timing of the air-fuel mixture is appropriate. An engine is disclosed that Specifically, in Patent Document 1, a pressure sensor that detects the pressure in the cylinder is provided, and after combustion is performed in the cylinder, heat is generated for each crank angle based on the change in the pressure in the cylinder during combustion. After that, an engine is disclosed in which the crank angle at which the amount of heat release is 50 to 60% of the maximum value is searched, and the ignition timing is adjusted so that this crank angle becomes the target value.

Further, in an engine provided in a vehicle or the like, execution of compression ignition combustion, in which an air-fuel mixture is self-ignited, is also being studied in order to improve fuel efficiency. When the air-fuel mixture is burned by self-ignition, the combustion period can be shortened compared to the case where the air-fuel mixture is ignited and burned by flame propagation, so the fuel consumption performance is improved. However, in the method of self-igniting the air-fuel mixture only by compression, there is a possibility that the air-fuel mixture may not self-ignite appropriately depending on the temperature state, the flow state of the gas in the cylinder, and the like. On the other hand, according to the method of performing partial compression ignition combustion in which the air-fuel mixture is ignited and part of the air-fuel mixture is burned by spark ignition, and then the remaining air-fuel mixture is burned by self-ignition, the temperature conditions and the like are different. It is believed that a portion of the air-fuel mixture can be burned by auto-ignition under a wide range of operating conditions.

JP 2010-127103 A

In the partial compression ignition combustion, ignition forces the combustion of the air-fuel mixture to start, and the generated combustion energy can raise the temperature of the air-fuel mixture, ensuring that the remaining air-fuel mixture self-ignites. However, the timing of self-ignition of the air-fuel mixture changes due to the difference in cylinder temperature immediately before ignition. Here, for example, if the self-ignition of the air-fuel mixture finally occurs at the timing when the piston descends significantly from the compression top dead center, the effect of improving the thermal efficiency and thus the fuel consumption performance is greatly reduced. On the other hand, using the configuration of Patent Document 1, the crank angle at which the amount of heat release becomes a predetermined ratio to the maximum value is searched, and the ignition timing is adjusted so that this crank angle becomes the target value. can be adjusted. However, in the configuration of Patent Document 1, the crank angle is obtained after combustion is completed, and the ignition timing is adjusted after one combustion cycle. There is a problem that the timing cannot be adjusted to a sufficiently appropriate timing.

SUMMARY OF THE INVENTION The present invention has been made in view of the circumstances as described above. The object is to provide a control device.

In order to solve the above-mentioned problems, the present invention comprises an engine body in which a cylinder is formed, and an ignition device that ignites the air-fuel mixture in the cylinder at a predetermined crank angle timing. A control method for a compression ignition engine capable of executing partial compression ignition combustion in which a part of the air-fuel mixture in the cylinder is burned by spark ignition and the remaining air-fuel mixture is burned by self-ignition, the method comprising: after the latter half of the compression stroke and an in-cylinder temperature estimating step of estimating a pre-ignition in-cylinder temperature, which is a temperature in the cylinder at a timing prior to the predetermined crank angle timing; The pre-ignition in-cylinder estimated in the in-cylinder temperature estimation step determines the crank angle period up to the index crank angle timing, which is the crank angle timing at which the combustion of a predetermined proportion of the fuel supplied to the engine is completed. An index crank angle period estimating step of estimating based on temperature and the crank angle period estimated in the index crank angle period estimating step, the index crank angle timing becomes a preset target index crank angle timing. and an ignition timing correcting step of correcting the predetermined crank angle timing such that, in the index crank angle period estimating step, the higher the pre-ignition in-cylinder temperature estimated in the in-cylinder temperature estimating step, the higher the pre-ignition in-cylinder temperature. An engine control method is provided, characterized in that the crank angle period from the predetermined crank angle timing to the index crank angle timing is estimated so as to be short.

According to the present invention, the index crank angle timing, which is the crank angle timing (timing with respect to the crank angle) at which a predetermined proportion of the fuel supplied to the cylinder during one combustion cycle completes combustion, is set. Predetermined crank angle timing, that is, ignition timing (timing for the crank angle at which ignition is performed) is corrected so as to achieve the target index crank angle timing, so the index crank angle timing can be made more appropriate. As a result, the partial compression ignition combustion can be performed at an appropriate timing regardless of changes in the temperature inside the cylinder, and the thermal efficiency and the fuel efficiency can be improved.

Moreover, in the present invention, the pre-ignition in-cylinder temperature, which is the temperature in the cylinder at the timing after the latter half of the compression stroke and before the predetermined crank angle timing, is estimated. Based on this, the crank angle period, which is the period (period of crank angle) from the predetermined crank angle timing to the index crank angle timing, is estimated. Based on the estimated crank angle period, the ignition timing is corrected so that the index crank angle timing becomes the target index crank angle timing. In other words, the index crank angle timing is estimated based on the temperature in the cylinder immediately before ignition, and the ignition timing is corrected so that it becomes the target value. Therefore, compared to the case where the index crank angle timing is calculated after the combustion of the air-fuel mixture is completed and the ignition timing is corrected so that it becomes the target value, the ignition timing can be corrected to an appropriate timing earlier. . For example, in a predetermined combustion cycle, before the combustion of the air-fuel mixture starts, it is also possible to correct the ignition timing so that the index crank angle timing of this combustion cycle becomes the target value.

Further, in the index crank angle period estimation step of the present invention , the higher the pre-ignition in-cylinder temperature estimated in the in-cylinder temperature estimation step, the higher the crank angle period from the predetermined crank angle timing to the index crank angle timing. The crank angle period is estimated so that is shortened .

In partial compression ignition combustion, the higher the temperature in the cylinder before ignition is performed, the earlier the air-fuel mixture will self-ignite and the more fuel will be burned at an earlier timing. , and it is known that the crank angle period, which is the period from the ignition timing to the index crank angle timing for the crank angle, is shortened. Therefore, according to this configuration, it is possible to accurately estimate the period from the ignition timing to the index crank angle timing.

In the above configuration, preferably, in the in-cylinder temperature estimating step, the pre-ignition in-cylinder temperature is estimated based on the temperature in the cylinder at the closing timing of an intake valve provided in the engine body. 2 ).

According to this configuration, the pre-ignition in-cylinder temperature can be accurately estimated. Moreover, the estimation of the pre-ignition in-cylinder temperature, and thus the index crank angle period estimation step and the ignition timing correction step can be performed sufficiently earlier than the ignition timing. Therefore, in a predetermined combustion cycle, before the combustion of the air-fuel mixture starts, the index crank angle timing of this combustion cycle can be controlled to the target index crank angle timing more reliably, and appropriate compression ignition can be performed more reliably. Combustion can be achieved.

The present invention comprises an engine body in which a cylinder is formed, and an ignition device that ignites an air-fuel mixture in the cylinder at a predetermined crank angle timing. A control device for a compression ignition engine capable of executing partial compression ignition combustion in which the remaining air-fuel mixture is burned by spark ignition combustion and self-ignition combustion, comprising control means for controlling the ignition device, wherein the control means an in-cylinder temperature estimating unit for estimating a pre-ignition in-cylinder temperature, which is the temperature in the cylinder after the latter half of the compression stroke and before the predetermined crank angle timing; The in-cylinder temperature estimator estimates a crank angle period up to an index crank angle timing, which is a crank angle timing at which a predetermined proportion of the fuel supplied to the cylinder during the combustion cycle completes combustion. an index crank angle period estimator for estimating based on the pre-ignition in-cylinder temperature that has been obtained; and the index crank angle timing is set in advance based on the crank angle period estimated by the index crank angle period estimator. an ignition timing correcting section that corrects the predetermined crank angle timing so as to achieve the target index crank angle timing, wherein the index crank angle period estimating section adjusts the ignition pre-cylinder temperature estimated by the in-cylinder temperature estimating section; An engine control device is provided, characterized in that the crank angle period is estimated so that the crank angle period from the predetermined crank angle timing to the index crank angle timing becomes shorter as the internal temperature increases (claim 3) . ).

With this device, similarly to the above-described method, a predetermined crank angle, that is, ignition timing (timing for the crank angle at which ignition is performed) is corrected so that the index crank angle timing becomes the target index crank angle timing. , appropriate partial compression ignition combustion can be realized to reliably improve fuel efficiency. In addition, the index crank angle timing is estimated based on the temperature in the cylinder before combustion starts, and the ignition timing is corrected so that it becomes the target index crank angle timing. can be corrected in time. In addition, in partial compression ignition combustion, the timing of self-ignition combustion of the air-fuel mixture can be accurately estimated, and the air-fuel mixture can be self-ignited and burned at more appropriate timing.

Further, according to this device , similarly to the above configuration, it is possible to accurately estimate the period from the ignition timing to the index crank angle timing, that is, the index crank angle period.

In the above configuration, preferably, the in-cylinder temperature estimating section estimates the pre-ignition in-cylinder temperature based on the temperature in the cylinder at the closing timing of an intake valve provided in the engine body. 4 ).

According to this configuration, similarly to the above configuration, the pre-ignition in-cylinder temperature can be accurately estimated, and in a predetermined combustion cycle, before the combustion of the air-fuel mixture starts, the combustion cycle can be more reliably performed. The ignition timing can be adjusted so that the index crank angle timing matches the target index crank angle timing, and more reliably appropriate compression ignition combustion can be achieved.

As described above, according to the engine control method and control device of the present invention, the air-fuel mixture can be combusted more reliably at an appropriate timing.

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. 2 is a map diagram in which engine operating regions are classified according to differences in combustion mode; 4 is a graph showing waveforms of heat release rate and heat release amount during SPCCI combustion (partial compression ignition combustion). FIG. 4 is a diagram schematically showing the valve lift of intake and exhaust valves, the injection pattern, and the heat release rate during SPCCI combustion. 4 is a flowchart showing a procedure for correcting ignition timing;

(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 . 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 typically 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 explanation will proceed by paying attention to only one cylinder 2 .

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. Hereinafter, the amount of fuel injected from the injector 15 is simply referred to as injection amount.

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 .

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 . By controlling these intake VVT 13a and exhaust VVT 14a, in this embodiment, it is possible to adjust the valve overlap period during which both the intake valve 11 and the exhaust valve 12 are opened across the exhaust top dead center. Further, by adjusting the valve overlap period, it is possible to adjust the amount of burned gas (internal EGR gas) remaining in the combustion chamber 6 . 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 spark plug 16 corresponds to the "igniter" 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 . Although illustration is omitted, in this embodiment, a cavity is formed in the crown surface of the piston 5 by recessing a region including the central portion thereof on 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 .

The intake passage 30 includes, in order from the upstream side thereof, an air cleaner 31 that removes foreign matter from the intake air, a throttle valve 32 that can be opened and closed to adjust the flow rate of the intake air, a supercharger 33 that compresses and delivers the intake air, and a supercharger. An intercooler 35 for cooling the intake air compressed by the feeder 33 and a surge tank 36 are provided.

Each portion of the intake passage 30 is provided with an airflow sensor SN4 for detecting the flow rate of intake air (intake amount) and an intake air temperature sensor SN5 for detecting the temperature of the intake air (intake air temperature). 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 amount of intake air that flows backward through the bypass passage 38 to the upstream side of the supercharger 33 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.

The exhaust passage 40 is provided with an exhaust temperature sensor SN6 that detects the temperature of the exhaust (exhaust temperature). The exhaust temperature sensor SN6 is provided in a portion of the exhaust passage 40 upstream 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 SN7 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 is composed of a well-known CPU, 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 coolant temperature sensor SN2, in-cylinder pressure sensor SN3, air flow sensor SN4, intake air temperature sensor SN5, exhaust temperature sensor SN6, and differential pressure sensor SN7. (i.e., crank angle, engine speed, engine water temperature, cylinder pressure, intake air amount, intake air temperature, exhaust temperature, differential pressure across EGR valve 53) detected by the sensors are sequentially input to ECU 100. there is 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 corresponds to "control means" in the claims. ECU 100 functionally includes in-cylinder temperature estimator 101 , index period estimator 102 , and ignition timing corrector 103 . The index period estimator 102 corresponds to the "index crank angle period estimator" in the claims.

(3) Combustion Control FIG. 3 is a map diagram for explaining differences in control according to engine speed and engine load. As shown in this figure, the operating range of the engine is broadly divided into two operating ranges, the SPCCI range A and the SI range B. FIG. The SPCCI region A is a region where the engine speed is low, and the SI region B is a region where the engine speed is high. ECU 100 determines whether the current operating point is included in SPCCI region A or SI region B based on the engine speed and engine load detected by crank angle sensor SN1, and performs the control described below. implement. The ECU 100 calculates the engine load based on the opening of the accelerator pedal detected by the accelerator sensor SN8, the engine speed, and the like.

(3-1) SPCCI Region (a) Combustion Mode In the SPCCI region A, partial compression ignition combustion (hereinafter referred to as SPCCI combustion) in which SI combustion and CI combustion are mixed is performed. "SPCCI" in SPCCI combustion is an abbreviation for "Spark Controlled Compression Ignition."

SI combustion is a form 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 expands the combustion area from the ignition point to the surroundings. 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.

(b) Contents of Control In the SPCCI region A, each part of the engine is controlled by the ECU 100 as follows so that SPCCI combustion is realized.

The injector 15 injects all or most of the fuel to be injected during one cycle during the intake stroke. For example, as shown in FIG. 5, in the SPCCI region A, most of the fuel is injected during the intake stroke (Q1), and the remaining fuel is injected twice during the compression stroke (Q2, Q3). . In the low-load region of the SPCCI region A, the air-fuel ratio (A/F) in the combustion chamber 6 is made higher (lean) than the stoichiometric air-fuel ratio. That is, the injector 15 injects into the combustion chamber 6 an amount of fuel that makes the air-fuel ratio (A/F) in the combustion chamber 6 higher than the stoichiometric air-fuel ratio. For example, in the low load region, the air-fuel ratio in the combustion chamber 6 is set to about 30. On the other hand, in the high-load region of the SPCCI region A, the air-fuel ratio in the combustion chamber 6 is close to the stoichiometric air-fuel ratio. That is, the injector 15 injects an amount of fuel into the combustion chamber 6 that makes the air-fuel ratio in the combustion chamber 6 near the stoichiometric air-fuel ratio.

The spark plug 16 ignites the air-fuel mixture in the vicinity of the compression top dead center in the entire SPCCI region A. 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.

The throttle valve 32 is fully opened in the entire SPCCI region A.

The EGR valve 53 is fully closed in the SPCCI region A on the low load side where the air-fuel ratio in the combustion chamber 6 is higher than the stoichiometric air-fuel ratio. As a result, in this region, introduction of the exhaust gas (external EGR gas) recirculated to the intake passage 30 through the EGR passage 51 into the combustion chamber 6 is stopped. On the other hand, the EGR valve 53 is opened in the high load region. However, in the high load region, the EGR valve 53 is controlled so that the amount of external EGR gas decreases as the load increases. is fully closed so that the amount of external EGR gas introduced into the combustion chamber 6 is substantially zero.

For the intake VVT 13a and the exhaust VVT 14a, the valve timings of the intake valve 11 and the exhaust valve 12 are set so that the internal EGR is performed only in the low load side (in other words, the internal EGR is stopped in the high load side). do. Specifically, in the low-load region of the SPCCI region A, as shown in FIG. The timing is set such that the valve overlap period in which the valve is opened across the dead center is sufficiently formed, and internal EGR that causes the burned gas to remain in the combustion chamber 6 is realized.

The swirl valve 18 is controlled such that its opening increases as the load increases. Specifically, in the low load region, the swirl valve 18 is fully closed or closed to a low degree of opening close to being fully closed. As a result, a strong swirl flow is formed in the combustion chamber 6 in the low load region. On the other hand, in the high load region, the swirl valve 18 is opened to an appropriate intermediate degree of opening other than fully closed/fully opened.

The supercharger 33 is turned off in the SPCCI region A where the load is low and the engine speed is low. 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. On the other hand, in other areas of the SPCCI area A, the supercharger 33 is turned on. That is, supercharging by the supercharger 33 is performed by engaging the electromagnetic clutch 34 and connecting 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).

(3-2) SI region B
In the SI region B, relatively orthodox SI combustion is performed. In order to realize this SI combustion, in the SI region B, each part of the engine is controlled by the ECU 100 as follows.

In the SI region B, 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 SI region B, this ignition triggers SI combustion, and all of the air-fuel mixture in the combustion chamber 6 is combusted by flame propagation.

The supercharger 33 is turned on. The throttle valve 32 is fully opened. The opening of the EGR valve 53 is controlled so that the air-fuel ratio (A/F) in the combustion chamber 6 is the stoichiometric air-fuel ratio or slightly richer than this. The swirl valve 18 is fully opened.

(4) Ignition Timing Control As described above, in SPCCI combustion, part of the air-fuel mixture undergoes CI combustion to shorten the combustion period and improve fuel efficiency. In addition, by igniting the air-fuel mixture and forcibly burning a part of the air-fuel mixture (SI combustion), the temperature of the air-fuel mixture is raised by the energy generated by this combustion, and the remaining air-fuel mixture is reliably CI-burned. can be made However, the temperature rise speed of the air-fuel mixture changes depending on the combustion speed of the SI combustion, and the timing at which self-ignition of the air-fuel mixture starts fluctuates. Therefore, if the timing at which the air-fuel mixture self-ignites is delayed and CI combustion, which is the main combustion, occurs at the timing when the piston 5 is significantly lowered from the compression top dead center, the effect of improving fuel efficiency cannot be sufficiently obtained. . That is, when the combustion period of SI combustion is lengthened as indicated by the chain line with respect to the heat release rate indicated by the solid line in FIG. As a result, the engine torque decreases. Therefore, in the present embodiment, the ignition timing, which is the timing at which the air-fuel mixture is ignited by the spark plug 16, is adjusted so that the CI combustion occurs at an appropriate timing in the SPCCI combustion.

As described above, the timing at which CI combustion, which is the main combustion of SPCCI combustion, occurs mainly changes according to the combustion speed of SI combustion. The combustion speed of SI combustion depends on the temperature of the air-fuel mixture at the start of SI combustion. Therefore, in the present embodiment, the temperature of the air-fuel mixture immediately before SI combustion starts, that is, the temperature in the combustion chamber 6 is estimated, and the ignition plug 16 ignites the air-fuel mixture based on this estimated value. Controls ignition timing. Here, in this specification, the ignition timing means the timing at which ignition is performed with respect to the crank angle, and the "predetermined crank angle timing" in the claims corresponds to this ignition timing.

The content of ignition timing control performed by the ECU 100 will be described with reference to the flowchart of FIG. 6 .

First, in step S1, the ECU 100 acquires values detected by the sensors SN1 to SN8.

Next, in step S2, the ECU 100 sets a basic ignition timing (basic value of ignition timing in terms of crank angle), which is a basic value of ignition timing. In this embodiment, the basic ignition timing is set in advance for the engine speed and the engine load and stored in the map in the ECU 100, and the ECU 100 extracts values corresponding to the current engine speed and the engine load from this map. do. Further, in step S2, the ECU 100 sets a target index timing, which is a target value of the index timing described later. Similar to the basic ignition timing, the target index timing is preset for the engine speed and the engine load and stored in the ECU 100 as a map. to extract

Next, in step S3, the ECU 100 estimates the in-cylinder temperature at the valve closing timing IVC of the intake valve 11 (hereinafter referred to as intake valve closing timing IVC). Here, as the valve closing timing IVC, any of the timing at which the intake valve 11 starts closing, the timing at which closing is completed, and an arbitrary timing between the start of closing and the completion of closing is adopted. can be The in-cylinder temperature is the temperature in the combustion chamber 6 , that is, the temperature of the air-fuel mixture in the combustion chamber 6 .

The gas in the combustion chamber 6 at the intake valve closing timing IVC is composed of the intake air (new) introduced into the combustion chamber 6, the external EGR gas introduced into the combustion chamber 6, and the burned gas remaining in the combustion chamber 6 after the exhaust stroke. The cylinder temperature is determined by the amount and temperature of these gases. Accordingly, the ECU 100 closes the intake valve based on the temperature and amount of the intake air, that is, the intake air amount, the temperature and amount of the external EGR gas, and the temperature and amount of the internal EGR gas, that is, the internal EGR gas amount. The in-cylinder temperature at timing IVC (hereinafter referred to as the in-cylinder temperature at intake valve closing) is estimated.

Specifically, the ECU 100 uses the value detected by the intake air temperature sensor SN5 at the intake valve closing timing IVC as the temperature of the intake air for estimating the in-cylinder temperature. The ECU 100 calculates the amount of intake air (intake amount) introduced into the combustion chamber 6 at the intake valve closing timing IVC based on the detection value of the airflow sensor SN4. The ECU 100 calculates the temperature of the external EGR gas and the temperature of the internal EGR gas introduced into the combustion chamber 6 at the intake valve closing timing IVC based on the detected value of the exhaust temperature sensor SN6. The ECU 100 calculates the amount of external EGR gas introduced into the combustion chamber 6 at the intake valve closing timing IVC based on the detected value of the differential pressure sensor SN7. The ECU 100 calculates the amount of internal EGR gas based on the detected value of the airflow sensor SN4 and the valve overlap period of the intake valve 11 and the exhaust valve 12 . Then, the ECU 100 uses these values to estimate the in-cylinder temperature when the intake valve is closed.

Next, in step S4, the ECU 100 determines the in-cylinder temperature (combustion The temperature of the air-fuel mixture in the chamber 6) is estimated. In this embodiment, the ECU 100 determines the in-cylinder temperature at a preset reference timing (crank angle) after the latter half of the compression stroke and before the basic ignition timing (hereinafter referred to as pre-ignition in-cylinder temperature). presume. The reference timing is set after the latter half of the compression stroke and before the basic ignition timing, for example, near the top dead center of the compression stroke. Steps S3 and S4 are performed by in-cylinder temperature estimating section 101 of ECU 100, and these steps S3 and S4 correspond to the "in-cylinder temperature estimating step" in the claims.

Here, when the intake valve 11 is closed at the intake valve closing timing IVC, the air-fuel mixture is compressed as the piston 5 rises, and the temperature rises. At this time, the higher the wall temperature of the combustion chamber 6, the smaller the cooling loss and the greater the temperature rise of the air-fuel mixture. That is, the higher the temperature of the engine cooling water detected by the engine water temperature sensor SN2 (hereinafter referred to as the engine water temperature), the greater the temperature rise of the air-fuel mixture. Also, the smaller the difference between the intake valve closing timing IVC and the intake bottom dead center BDC (for example, the smaller the retardation amount of the intake valve closing timing IVC from the intake bottom dead center BDC), the larger the effective compression ratio. As a result, the amount of compression by the piston 5 until the reference timing increases, so the amount of temperature rise of the air-fuel mixture increases. It is also known that the lower the engine speed, the greater the temperature rise of the air-fuel mixture due to the compression by the piston 5 . From this, the ECU 100 estimates the pre-ignition in-cylinder temperature based on the in-cylinder temperature when the intake valve is closed, the engine coolant temperature, the intake valve-closing timing IVC, ie, the effective compression ratio, and the engine speed, as described above. Specifically, the ECU 100 determines that the higher the intake valve closing time cylinder temperature, the higher the engine water temperature, the smaller the difference between the intake valve closing timing IVC and the intake bottom dead center BDC, and the lower the engine speed, This is estimated so that the pre-ignition in-cylinder temperature increases.

Next, in step S5, the ECU 100 compares the pre-ignition in-cylinder temperature estimated in step S4 with a preset reference temperature, and based on the comparison result, the basic ignition timing set in step S2 is used as an index. Estimate the indicator period, which is the period to time. This step S5 is performed by the index period estimating section 102 of the ECU 100, and corresponds to the "index crank angle period estimating step" in the claims.

The index timing is the timing at which the combustion of a predetermined proportion (mass) of fuel out of the total fuel (mass) supplied to the combustion chamber 6 during one combustion cycle is completed. In this embodiment, the predetermined ratio is set to 50%, and the so-called combustion center of gravity timing, which is the timing at which 50% of the total fuel supplied to the combustion chamber 6 during one combustion cycle completes combustion, is used as an index. It is time. This is because in SPCCI combustion, the timing of the center of gravity of combustion is included in the period during which CI combustion is being performed. As shown in the graph of the amount of heat release in FIG. 4, the period from the ignition timing to the timing of the center of gravity of combustion is used as the index period, and in step S5, the period from this timing of ignition to the timing of the center of gravity of combustion is estimated.

Here, the index timing referred to in this specification is the timing of the crank angle, and the amount (mass) of a predetermined ratio of the total fuel (mass) supplied to the combustion chamber 6 during one combustion cycle. The crank angle at which combustion is completed. Similarly, the timing of the center of gravity of combustion also refers to the timing of the crank angle. This index timing corresponds to the "index crank angle timing" in the claims. Further, the index period referred to in this specification is a period regarding the crank angle, and refers to a period regarding the crank angle from the ignition timing to the index timing (timing of the center of gravity of combustion).

The higher the in-cylinder temperature immediately before ignition, the faster the combustion speed of SI combustion, and the earlier the timing at which CI combustion starts and the timing at the center of gravity of combustion. Accordingly, in step S5, the ECU 100 estimates the index period such that the higher the pre-ignition in-cylinder temperature estimated in step S4, the shorter the index period.

Specifically, a reference period, which is an index period when the pre-ignition in-cylinder temperature is at a predetermined reference temperature, is obtained in advance by experiments or the like and stored in the ECU 100 . The ECU 100 compares the pre-ignition in-cylinder temperature estimated in step S4 with the reference temperature, and when the pre-ignition in-cylinder temperature is the reference temperature, the ECU 100 estimates the reference period as the index period. When the temperature is higher than the reference period, a period shorter than the reference period is estimated as the index period, and when the in-cylinder temperature before ignition is lower than the reference temperature, a period longer than the reference period is estimated as the index period. In this embodiment, a plurality of reference temperatures and reference periods are determined in advance for each engine speed and each engine load and stored in the ECU 100 . The ECU 100 extracts the reference temperature and the reference period corresponding to the current engine speed and engine load, and compares them with the pre-ignition in-cylinder temperature estimated in step S4. Further, the amount of deviation of the in-cylinder temperature before ignition from the reference temperature and the amount of deviation of the reference period of the index period when this deviation occurs are preset and stored in the ECU 100 as a map. The index period is estimated by calculating the amount of deviation between the in-cylinder temperature and the reference temperature, and extracting the corresponding amount of deviation in the index period from the map.

Next, in step S6, the ECU 100 determines the ignition timing by correcting the basic ignition timing set in step S2 based on the index period estimated in step S5. At this time, the ECU 100 corrects the ignition timing so that the index timing, that is, the combustion center-of-gravity timing, becomes the target value of the target index timing, that is, the combustion center-of-gravity timing set in step S2. Specifically, the ECU 100 determines the ignition timing to be the basic ignition timing when the index period estimated in step S5 is the same as the reference period or when it can be considered that there is substantially no difference. When the index period is longer than the reference period, the ignition timing is advanced from the basic ignition timing, and when the index period is shorter than the reference period, the ignition timing is retarded from the basic ignition timing. For example , the basic ignition timing is corrected so that the ignition timing is advanced by the index period from the target index timing. This step S6 is performed by the ignition timing correcting section 103 of the ECU 100, and corresponds to the "ignition timing correcting step" in the claims.

Next, in step S7, the ECU 100 controls the spark plug 16 so that ignition is performed at the ignition timing set in step S6.

(5) Operation, etc. As described above, in the present embodiment, the ignition timing is corrected so that the combustion center-of-gravity timing becomes the target index timing. Therefore, even if the combustion speed of SI combustion changes due to temperature changes in the combustion chamber 6 before ignition, the air-fuel mixture can be CI-burned at more appropriate timing. Therefore, the fuel consumption performance can be improved more reliably.

In particular, in the present embodiment, the pre-ignition in-cylinder temperature, which is the temperature in the combustion chamber 6 at the timing after the latter half of the compression stroke and before the ignition timing, is estimated, and based on the estimated pre-ignition in-cylinder temperature. Then, the index period, which is the timing from the ignition timing to the combustion center of gravity timing, is estimated. That is, the combustion center-of-gravity timing is estimated based on the temperature in the combustion chamber 6 immediately before combustion starts, and the ignition timing for that cycle is corrected based on this. Therefore, compared to the case where the combustion center of gravity timing is calculated after the combustion of the air-fuel mixture ends and the ignition timing of the next cycle or subsequent cylinder is corrected based on this, the ignition timing is corrected to the appropriate timing earlier. be able to. Therefore, since CI combustion can be started at an appropriate timing, it is possible to avoid the occurrence of misfires and excessive increase in combustion noise due to SPCCI combustion in which the timing of the center of gravity of combustion deviates from the target index timing.

Further, as described above, in SPCCI combustion, the higher the pre-ignition in-cylinder temperature, the higher the combustion speed of SI combustion and the earlier CI combustion starts. From this, the higher the pre-ignition in-cylinder temperature, the earlier the timing of the combustion center of gravity and the shorter the index period. In contrast, in the present embodiment, the index period is estimated such that the higher the pre-ignition in-cylinder temperature, the shorter the index period. Therefore, the index period and the combustion center-of-gravity timing can be estimated more accurately.

Further, in the present embodiment, the pre-ignition in-cylinder temperature is estimated based on the in-cylinder temperature at intake valve closing, which is the temperature in the combustion chamber 6 at the intake valve closing timing IVC. can be estimated. Specifically, the air-fuel mixture that exists and burns in the combustion chamber 6 in the latter half of the compression stroke is the air-fuel mixture that exists in the combustion chamber 6 after the intake valve 11 closes, and burns at the intake valve closing timing IVC. The temperature resulting from the air-fuel mixture in the chamber 6 being compressed and raised is the pre-ignition in-cylinder temperature. Therefore, by estimating the pre-ignition in-cylinder temperature based on the intake valve closed time in-cylinder temperature, it is possible to estimate the pre-ignition in-cylinder temperature with higher accuracy at a timing that can be in time for the ignition of the cycle. Furthermore, by estimating the pre-ignition in-cylinder temperature based on the temperature in the combustion chamber 6 at the intake valve closing timing IVC, the index period can be estimated and the ignition timing can be corrected at a timing close to the intake valve closing timing IVC. can be done. Therefore, before the combustion of the air-fuel mixture starts in a predetermined combustion cycle, the ignition timing of this combustion cycle can be more reliably corrected to an appropriate timing.

(6) Modification In the above-described embodiment, the index timing used for controlling the ignition timing is a predetermined proportion (mass) of the total fuel (mass) supplied to the combustion chamber 6 during one combustion cycle. The timing of completing the combustion of is the combustion center-of-gravity timing, that is, the case where the predetermined ratio is 50%, but the predetermined ratio is not limited to 50% as long as it is greater than 0%. However, as described above, the timing of the center of gravity of combustion is included in the implementation period of CI combustion. Therefore, if the index timing is set as the combustion center of gravity timing and the ignition timing is controlled so that this index timing becomes the target index timing, the combustion timing of the CI combustion can be made more appropriate, and the fuel efficiency performance can be effectively improved. can be enhanced.

In the above embodiment, the case of estimating the pre-ignition in-cylinder temperature using the in-cylinder temperature at the intake valve closing timing IVC (the in-cylinder temperature at the time of intake valve closing) has been described. The internal temperature is not limited to the in-cylinder temperature at the intake valve closing timing IVC.

Further, in the above embodiment, a case has been described in which the in-cylinder temperature at the intake valve closing timing IVC is estimated, and the pre-ignition in-cylinder temperature is estimated based on this estimated value. On the other hand, the temperature detected by the intake air temperature sensor SN5 at the intake valve closing timing IVC may be used as the in-cylinder temperature at the intake valve closing timing IVC.

In the above embodiment, the index period is estimated based on the pre-ignition in-cylinder temperature, and based on this estimated value, the control for correcting the ignition timing so that the index timing becomes the target index timing is performed in the SPCCI region A. As described, this control may be performed in the SI region B as well.

Reference Signs List 1 engine body 2 cylinder 6 combustion chamber 15 injector 16 spark plug 41a three-way catalyst 100 ECU (control means)
101 in-cylinder temperature estimator 102 index period estimator (index crank period estimator)
103 Ignition timing corrector

Claims (4)

An engine body having a cylinder and an ignition device for igniting an air-fuel mixture in the cylinder at a predetermined crank angle timing. A control method for a compression ignition engine capable of executing partial compression ignition combustion in which the mixture is burned and the remaining air-fuel mixture is burned by self-ignition,
an in-cylinder temperature estimating step of estimating a pre-ignition in-cylinder temperature, which is the temperature in the cylinder after the latter half of the compression stroke and before the predetermined crank angle timing;
A crank angle period from the predetermined crank angle timing to an index crank angle timing, which is a crank angle timing at which a predetermined proportion of the fuel supplied to the cylinder in one combustion cycle completes combustion. an index crank angle period estimation step of estimating based on the pre-ignition in-cylinder temperature estimated in the in-cylinder temperature estimation step;
ignition timing correction for correcting the predetermined crank angle timing so that the index crank angle timing becomes a preset target index crank angle timing based on the crank angle period estimated in the index crank angle period estimating step; and
In the index crank angle period estimation step, the higher the pre-ignition in-cylinder temperature estimated in the in-cylinder temperature estimation step, the shorter the crank angle period from the predetermined crank angle timing to the index crank angle timing. and estimating the crank angle period at a time . In the engine control method according to claim 1 ,
In the in-cylinder temperature estimating step, the pre-ignition in-cylinder temperature is estimated based on the temperature in the cylinder at the closing timing of an intake valve provided in the engine body. . An engine body having a cylinder and an ignition device for igniting an air-fuel mixture in the cylinder at a predetermined crank angle timing. A control device for a compression ignition engine capable of executing partial compression ignition combustion in which the remaining air-fuel mixture is burned by self-ignition,
A control means for controlling the ignition device,
The control means is
an in-cylinder temperature estimating unit for estimating a pre-ignition in-cylinder temperature, which is the temperature in the cylinder after the latter half of the compression stroke and before the predetermined crank angle timing;
A crank angle period from the predetermined crank angle timing to an index crank angle timing, which is a crank angle timing at which a predetermined proportion of the fuel supplied to the cylinder during one combustion cycle completes combustion. an index crank angle period estimator for estimating based on the pre-ignition in-cylinder temperature estimated by the in-cylinder temperature estimator;
ignition timing correction for correcting the predetermined crank angle timing based on the crank angle period estimated by the index crank angle period estimator so that the index crank angle timing becomes a preset target index crank angle timing; and
The index crank angle period estimating section is arranged such that the higher the pre-ignition in-cylinder temperature estimated by the in-cylinder temperature estimating section, the shorter the crank angle period from the predetermined crank angle timing to the index crank angle timing. An engine control device, characterized by estimating the crank angle period . In the engine control device according to claim 3 ,
The engine control device, wherein the in-cylinder temperature estimating unit estimates the pre-ignition in-cylinder temperature based on the in-cylinder temperature at the closing timing of an intake valve provided in the engine body. .

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