JP7263906B2 - engine controller - Google Patents

Description

The technology disclosed herein relates to an engine control device.

Patent Literature 1 describes a technique for estimating the temperature of exhaust gas in a spark ignition engine. Specifically, in this technology, the electronic control unit uses an exhaust temperature map that defines the relationship between the load and speed of the engine and the temperature of the exhaust gas, and determines the temperature of the exhaust gas based on the operating state of the engine. presume. The exhaust gas temperature is affected by the ignition timing. Therefore, Patent Document 1 describes that the estimated exhaust gas temperature is corrected by a correction coefficient determined according to the difference between the combustion center of gravity when the ignition timing is MBT and the combustion center of gravity of actual combustion. It is The correction coefficient is set to increase quadratically as the difference in the combustion center of gravity increases.

Japanese Patent Application Laid-Open No. 2017-223138

The applicant of the present application has proposed SPCCI (SPark Controlled Compression Ignition) combustion, which is a combination of SI (Spark Ignition) combustion and CI (Compression Ignition) combustion. SI combustion is combustion accompanied by flame propagation initiated by forcibly igniting the air-fuel mixture in the cylinder. CI combustion is combustion initiated by self-ignition of an air-fuel mixture in a cylinder. In SPCCI combustion, when the air-fuel mixture in the cylinder is forcibly ignited to start combustion due to flame propagation, the heat generated by SI combustion and the pressure increase due to flame propagation cause the unburned air-fuel mixture in the cylinder to self-ignite. It is in a burning form.

Even in an engine that switches between the SI combustion mode and the SPCCI combustion mode depending on the operating state, there is a demand to accurately estimate the temperature of the exhaust gas for the purpose of ensuring the reliability of the catalyst device, for example. However, no technology has yet been established for accurately estimating the temperature of the exhaust gas during execution of SPCCI combustion.

The technology disclosed herein accurately estimates the temperature of exhaust gas in an engine that switches between combustion modes.

The technology disclosed herein relates to an engine control device. This control device
engine and
an exhaust passage that is connected to the engine and guides exhaust gas from a cylinder of the engine;
a sensor that outputs a signal corresponding to the combustion state in the cylinder;
a control unit connected to the sensor, estimating the temperature of the exhaust gas based on the signal of the sensor, and controlling the engine according to the estimated temperature of the exhaust gas,
The control unit performs forced ignition of the air-fuel mixture in the cylinder and burns the air-fuel mixture by flame propagation according to the operating state of the engine. and switching between a second combustion mode in which part of the air-fuel mixture is burned by self-ignition,
The control unit
a calculation unit that calculates a combustion progress degree, which is a crank angle at which a mass combustion ratio of combustion in the cylinder is 10%, 50%, or 90%, based on the signal from the sensor;
an estimation unit that estimates the temperature of the exhaust gas from the degree of combustion progress calculated by the calculation unit based on the relationship between the degree of combustion progress and the temperature of the exhaust gas;
The estimation unit estimates the temperature of the exhaust gas based on a first relationship between the degree of combustion progress and the temperature of the exhaust gas during the first combustion mode, and estimates the temperature of the exhaust gas during the second combustion mode. estimating the temperature of the exhaust gas based on a second relationship between the degree of combustion and the temperature of the exhaust gas, different from the first relationship;

The control unit switches between the first combustion mode and the second combustion mode according to the operating state of the engine. The first combustion mode is a mode in which forced ignition is performed on the air-fuel mixture in the cylinder and the air-fuel mixture is burned by flame propagation. The first combustion mode corresponds to a so-called SI combustion mode. The second combustion mode is a mode in which the air-fuel mixture in the cylinder is forcibly ignited and part of the air-fuel mixture is burned by self-ignition. The second combustion mode corresponds to a so-called SPCCI combustion mode.

The engine control device having this configuration estimates the temperature of the exhaust gas based on the degree of progress of combustion. Specifically, the calculation unit of the control unit calculates the degree of combustion progress, which is the crank angle when the combustion in the cylinder has progressed to a specific degree, based on the signal from the sensor. The sensor outputs a signal corresponding to the combustion state in the cylinder. The sensor may be an in-cylinder pressure sensor that outputs a signal corresponding to the pressure in the cylinder. Based on the signal from the in-cylinder pressure sensor, the control unit can accurately grasp the combustion state in the cylinder.

Since the degree of combustion progress is the crank angle when combustion has progressed to a specific degree, it can be used as a parameter representing the state of combustion. A crank angle at which the mass fraction burned is a specific value, for example, a crank angle at which the mass fraction burned 50% (that is, mass fraction burned 50:mfb50) may be used as the degree of combustion progress. mfb50 means the crank angle at which 50% of the total injection amount of fuel is burned. Incidentally, the degree of combustion progress is not limited to mfb50. Any value such as mfb10 or mfb90 can be used as the degree of combustion progress if there is a correlation with the exhaust gas temperature.

The estimation unit of the control unit estimates the temperature of the exhaust gas from the degree of combustion progress calculated by the calculation unit based on the relationship between the degree of combustion progress and the temperature of the exhaust gas. The estimation unit also estimates the temperature of the exhaust gas based on the first relationship between the degree of combustion progress and the temperature of the exhaust gas during the first combustion mode, and estimates the temperature of the exhaust gas based on the first relationship during the second combustion mode. Estimate the temperature of the exhaust gas based on a different second relationship. The controller estimates the temperature of the exhaust gas based on different relationships between the degree of combustion progress and the temperature of the exhaust gas corresponding to different combustion modes. As a result, the control section can accurately estimate the temperature of the exhaust gas in the first combustion mode and the second combustion mode.

The estimator may estimate the temperature of the exhaust gas to be higher in the first combustion mode than in the second combustion mode when the degree of combustion progress is the same.

As described above, the second combustion mode burns part of the unburned air-fuel mixture by self-ignition, and therefore has higher thermal efficiency than the first combustion mode. Therefore, in the second combustion mode, the temperature of the exhaust gas is lower than in the first combustion mode, and conversely, in the first combustion mode, the temperature of the exhaust gas is higher than in the second combustion mode. The estimating unit estimates the temperature of the exhaust gas to be higher in the first combustion mode than in the second combustion mode for the same degree of combustion progress, so that the control unit accurately estimates the temperature of the exhaust gas in the first combustion mode. In addition, the temperature of the exhaust gas can be accurately estimated during the second combustion mode.

In each of the first combustion mode and the second combustion mode, when the degree of combustion progress is retarded relative to the degree of combustion progress in the MBT, the estimating unit calculates a , the temperature of the exhaust gas may be linearly estimated higher.

Here, MBT (Minimum advance for the Best Torque) is the ignition timing at which the torque of the engine becomes maximum when the ignition timing is changed while the driving state is fixed.

The inventors of the present application found that when the degree of combustion progress in SI combustion and SPCCI combustion lags behind the degree of combustion progress in MBT, there is a linear correlation between the degree of combustion progress and the temperature of the exhaust gas. found to exist.

When the ignition timing is retarded, the temperature of the exhaust gas rises. A correlation also exists between ignition timing and exhaust gas temperature. However, according to studies by the inventors of the present application, in engines that perform SI combustion and/or SPCCI combustion, in other words, engines that perform forced ignition, the relationship between the ignition timing and the temperature of the exhaust gas is non-linear rather than linear. there were.

With late ignition timing, combustion begins in the expansion stroke. The expansion stroke is a stroke in which the volume inside the cylinder is expanded. Volume expansion has a non-linear relationship to crank angle progression. Combustion in the expansion stroke is subject to non-linear volumetric expansion. Therefore, the degree of progress of combustion, which indicates the state of combustion in the cylinder, has a nonlinear relationship (for example, a quadratic function) with respect to the retardation of ignition timing.

Since the degree of combustion progress is a parameter representing the state of combustion, there is a linear correlation between the degree of combustion progress and the indicated work of the engine. Also, the temperature of the exhaust gas is determined by the heat amount obtained by subtracting the indicated work of the engine from the heat amount generated by combustion in the cylinder. Therefore, a linear correlation is also established between the indicated work of the engine and the temperature of the exhaust gas. As described above, the ignition timing and the degree of combustion progress have a non-linear relationship, so the correlation between the ignition timing and the temperature of the exhaust gas is non-linear. The correlation becomes linear.

A first relationship that is the correlation between the degree of combustion progress and the temperature of the exhaust gas during the first combustion mode, and a second correlation that is the correlation between the degree of combustion progress and the temperature of the exhaust gas during the second combustion mode. Both relationships are linear. In the first combustion mode, when the degree of combustion progresses more retarded than the degree of combustion progress in the MBT, the estimator linearly adjusts the temperature of the exhaust gas as the degree of combustion progresses more retarded based on the first relationship. estimated to be high. In the second combustion mode, when the degree of combustion progresses more retarded than the degree of combustion progress in the MBT, the estimating unit linearly adjusts the temperature of the exhaust gas as the degree of combustion progresses more retarded based on the second relationship. Estimate high. The estimator can accurately estimate the temperature of the exhaust gas during the first combustion mode and during the second combustion mode. Since the first relationship and the second relationship are linear, the estimator can accurately estimate the temperature of the exhaust gas even when the degree of combustion lags significantly behind the degree of combustion in the MBT. .

The estimating unit retards the degree of combustion progress when the degree of progress of combustion is retarded relative to the degree of progress of combustion at a knock limit in each of the first combustion mode and the second combustion mode. The temperature of the exhaust gas may be linearly estimated to be higher as the temperature increases.

Here, the knock limit is the ignition timing advance limit at which abnormal combustion does not occur. The knock limit is more retarded than MBT.

A linear correlation is established between the degree of combustion progress in SI combustion and SPCCI combustion and the temperature of the exhaust gas even when the degree of combustion lags behind the degree of combustion progress at the knock limit. The estimator can accurately estimate the temperature of the exhaust gas in the first combustion mode and the second combustion mode based on the signal from the sensor and the first relationship or the second relationship.

The second relationship may be such that the temperature rise rate of the exhaust gas with respect to changes in the degree of combustion progress is higher than the temperature rise rate of the first relationship.

In the second combustion mode, SPCCI combustion is performed. A portion of the air-fuel mixture is CI-burned. CI combustion performed in the expansion stroke is greatly affected by changes in volume. Specifically, when the degree of combustion progress is retarded, the thermal efficiency of SPCCI combustion is greatly reduced. As a result, during the second combustion mode, the temperature of the exhaust gas changes greatly with respect to changes in the degree of combustion progress.

In the first combustion mode, SI combustion is performed. Combustion by flame propagation is less sensitive to volumetric changes. Even if the degree of combustion progresses in the first combustion mode, the thermal efficiency of SI combustion does not significantly decrease. During the first combustion mode, changes in exhaust gas temperature are small with respect to changes in the degree of combustion progress.

Therefore, in the second relationship, the temperature rise rate of the exhaust gas with respect to the change in the degree of combustion progress is higher than in the first relationship. The control unit can accurately estimate the temperature of the exhaust gas in the first combustion mode and the second combustion mode based on the first relationship and the second relationship.

In each of the first combustion mode and the second combustion mode, the estimating unit calculates the temperature of the exhaust gas for the same degree of combustion progress when the engine speed is high as compared to when the engine speed is low. Estimate high.

When the engine speed is high, the temperature of the exhaust gas is higher than when the engine speed is low. This is because the higher the engine speed, the higher the cylinder wall temperature due to the increased number of combustions per unit time, and the shorter the combustion time in the cylinder, which reduces the cooling loss.

In each of the first combustion mode and the second combustion mode, the estimating unit estimates the exhaust gas temperature for the same degree of combustion progress to be higher when the engine speed is high than when the engine speed is low. . The control unit can accurately estimate the temperature of the exhaust gas even if the engine speed changes.

In each of the first relationship and the second relationship, the temperature rise rate of the exhaust gas with respect to changes in the degree of combustion progress may be higher when the engine speed is high than when the engine speed is low.

As described above, when the engine speed is high, the cylinder wall temperature is high, and the difference between the in-cylinder temperature and the cylinder wall temperature during combustion is relatively small. When the engine speed is high and the in-cylinder temperature changes as the degree of combustion progresses, the ratio of the amount of change in the in-cylinder temperature to the difference between the in-cylinder temperature and the cylinder wall temperature increases. That is, if the degree of combustion progresses when the engine speed is high, the cooling loss will change significantly, and the temperature of the exhaust gas will also change significantly. Conversely, when the engine speed is low, the cylinder wall temperature is low, and the difference between the in-cylinder temperature during combustion and the cylinder wall temperature is relatively large. When the engine speed is low, even if the in-cylinder temperature changes as the degree of combustion progresses, the ratio of the amount of change in the in-cylinder temperature to the difference between the in-cylinder temperature and the cylinder wall temperature is small. Gas temperature change is small.

The estimating unit estimates the temperature of the exhaust gas based on the fact that the temperature rise rate of the exhaust gas with respect to the change in the degree of combustion is higher when the engine speed is high than when the engine speed is low. The part can accurately estimate the temperature of the exhaust gas in the first combustion mode and in the second combustion mode even if the engine speed changes.

The estimating unit estimates the temperature of the exhaust gas for the same degree of combustion progress in each of the first combustion mode and the second combustion mode to be higher when the load of the engine is high than when the load is low. You can do it.

When the engine load is high, the amount of fuel supplied to the cylinder is large, so the amount of heat generated in the cylinder is large. Due to the large amount of heat, the temperature of the exhaust gas is also high. In each of the first combustion mode and the second combustion mode, the estimator estimates that the exhaust gas temperature for the same degree of combustion progress is higher when the engine load is high than when the engine load is low. This allows the control unit to accurately estimate the temperature of the exhaust gas even if the engine load changes.

In each of the first relationship and the second relationship, the temperature rise rate of the exhaust gas with respect to changes in the degree of combustion progress may be higher when the load of the engine is high than when the load is low.

As mentioned above, when the engine load is high, the amount of fuel supplied is large. When the engine load is high and the degree of combustion progresses, the amount of change in indicated work is large. As a result, the temperature of the exhaust gas also changes significantly. Conversely, when the engine load is low, the amount of fuel supplied is small, so even if the progress of combustion changes, the amount of change in indicated work is small. As a result, the change in temperature of the exhaust gas is small.

When the engine load is high, the temperature rise rate of the exhaust gas with respect to the change in the degree of combustion is higher than when the load is low. , the temperature of the exhaust gas can be accurately estimated in the first combustion mode and in the second combustion mode even if the engine load changes.

When the temperature of the exhaust gas is higher than a reference, the control unit may increase the amount of fuel supplied to the cylinder more than when the temperature is equal to or lower than the reference.

When the amount of fuel supplied into the cylinder is increased, the latent heat of the increased amount of fuel lowers the temperature of the exhaust gas discharged from the cylinder. By lowering the temperature of the exhaust gas below the standard, it is possible to ensure the reliability of the catalyst device interposed in the exhaust passage of the engine, for example. The performance of the catalyst device can be maintained at a high level, and the engine can exhaust cleaned gas.

As described above, the engine control device can accurately estimate the temperature of exhaust gas in an engine that switches between the first combustion mode and the second combustion mode.

FIG. 1 is a diagram illustrating the configuration of an engine. FIG. 2 is a block diagram illustrating the configuration of an engine control device. The upper diagram in FIG. 3 illustrates the waveform of SPCCI combustion, and the lower diagram illustrates the output signal of the crank angle sensor. FIG. 4 is a diagram illustrating an engine operation map. FIG. 5 is a diagram illustrating fuel injection timing, ignition timing, and combustion waveform in each operating region of the map of FIG. FIG. 6 is a block diagram illustrating functional blocks of an ECU that estimates the temperature of exhaust gas. FIG. 7 is a flowchart illustrating a procedure for estimating the temperature of the exhaust gas executed by the ECU. FIG. 8 is a diagram illustrating a model representing the relationship between the degree of combustion progress and the temperature of the exhaust gas. The upper diagram in FIG. 9 illustrates a model for SI combustion at a high engine speed and a model for low engine speed, and a lower diagram illustrates a model for SPCCI combustion at a high engine speed. FIG. 10 is a diagram illustrating a model for time and a model for low rotation; FIG. 10 is a diagram illustrating the relationship between engine speed and exhaust gas temperature. The upper diagram of FIG. 11 is a diagram illustrating a model at the time of SI combustion with a high engine load and a model at the time of a low load. FIG. 4 is a diagram illustrating a model and a model at low load; FIG. 12 is a diagram illustrating the relationship between the temperature of the coolant and the amount of correction for the temperature of the exhaust gas.

An embodiment of an engine control device will be described in detail below with reference to the drawings. The following description is an example of an engine, an engine system, and a controller for the engine.

FIG. 1 is a diagram illustrating the configuration of an engine system. FIG. 2 is a block diagram illustrating the configuration of an engine control device.

The engine 1 is a four-stroke engine that operates by repeating an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke in a combustion chamber 17 . The engine 1 is installed in a four-wheeled automobile. The automobile runs as the engine 1 operates. The fuel of the engine 1 is gasoline in the engine of this configuration example. The fuel may be liquid fuel containing at least gasoline. The fuel may be gasoline, including, for example, bioethanol.

(Engine configuration)
The engine 1 has a cylinder block 12 and a cylinder head 13 mounted thereon. A plurality of cylinders 11 are formed inside the cylinder block 12 . In FIG. 1 only one cylinder 11 is shown. The engine 1 is a multi-cylinder engine.

A piston 3 is slidably inserted in each cylinder 11 . Piston 3 is connected to crankshaft 15 via connecting rod 14 . Piston 3 defines combustion chamber 17 together with cylinder 11 and cylinder head 13 . Incidentally, the term "combustion chamber" may be used in a broad sense. In other words, the "combustion chamber" may mean the space formed by the piston 3, the cylinder 11 and the cylinder head 13 regardless of the position of the piston 3.

The geometric compression ratio of the engine 1 is set to 10 or more and 30 or less. As will be described later, the engine 1 performs SPCCI combustion, which is a combination of SI combustion and CI combustion, in some operating ranges. SPCCI combustion utilizes the heat generation and pressure increase due to SI combustion to control CI combustion. The engine 1 is of a compression ignition type. This engine 1 does not need to increase the temperature of the combustion chamber 17 (that is, the compression end temperature) when the piston 3 reaches the compression top dead center. The engine 1 can have a relatively low geometric compression ratio. A lower geometric compression ratio favors lower cooling losses and lower mechanical losses. The geometric compression ratio of the engine 1 is 14 to 17 for a regular specification engine (low octane fuel with an octane rating of about 91), and 14 to 17 for a high octane specification engine (high octane fuel with an octane rating of about 96). , 15-18.

An intake port 18 is formed in the cylinder head 13 for each cylinder 11 . Although not shown, the intake port 18 has a first intake port and a second intake port. The intake port 18 communicates with the combustion chamber 17 . Although detailed illustration is omitted, the intake port 18 is a so-called tumble port. That is, the intake port 18 has a shape that forms a tumble flow in the combustion chamber 17 .

An intake valve 21 is arranged in the intake port 18 . The intake valve 21 opens and closes between the combustion chamber 17 and the intake port 18 . The intake valve 21 is opened and closed at a predetermined timing by a valve mechanism. The valve mechanism may be a variable valve mechanism that varies valve timing and/or valve lift. In this configuration example, the variable valve mechanism has an electric intake S-VT (Sequential-Valve Timing) 23, as shown in FIG. The electric intake S-VT 23 continuously changes the rotation phase of the intake camshaft within a predetermined angle range. The intake valve mechanism may have a hydraulic S-VT instead of the electric S-VT.

The cylinder head 13 is also formed with an exhaust port 19 for each cylinder 11 . The exhaust port 19 also has a first exhaust port and a second exhaust port. The exhaust port 19 communicates with the combustion chamber 17 .

An exhaust valve 22 is arranged in the exhaust port 19 . The exhaust valve 22 opens and closes between the combustion chamber 17 and the exhaust port 19 . The exhaust valve 22 is opened and closed at a predetermined timing by a valve mechanism. This valve mechanism may be a variable valve mechanism that varies valve timing and/or valve lift. In this configuration example, the variable valve mechanism has an electric exhaust S-VT 24, as shown in FIG. The electric exhaust S-VT 24 continuously changes the rotation phase of the exhaust camshaft within a predetermined angle range. The exhaust valve mechanism may have a hydraulic S-VT instead of the electric S-VT.

The electric intake S-VT 23 and the electric exhaust S-VT 24 adjust the length of the overlap period during which both the intake valve 21 and the exhaust valve 22 are open. By increasing the length of the overlap period, residual gases in the combustion chamber 17 can be scavenged. Also, internal EGR (Exhaust Gas Recirculation) gas can be introduced into the combustion chamber 17 by adjusting the length of the overlap period. The electric intake S-VT 23 and the electric exhaust S-VT 24 constitute an internal EGR system. Note that the internal EGR system is not limited to the S-VT.

An injector 6 is attached to the cylinder head 13 for each cylinder 11 . Injector 6 injects fuel directly into combustion chamber 17 . The injector 6 is an example of a fuel supply section. Although detailed illustration is omitted, the injector 6 is configured by a multi-orifice fuel injection valve having a plurality of orifices. The injector 6 is attached to the center of the combustion chamber 17 and injects fuel so that the fuel spray spreads radially from the center of the combustion chamber 17 .

A fuel supply system 61 is connected to the injector 6 . The fuel supply system 61 includes a fuel tank 63 configured to store fuel, and a fuel supply path 62 connecting the fuel tank 63 and the injector 6 to each other. A fuel pump 65 and a common rail 64 are interposed in the fuel supply path 62 . Fuel pump 65 delivers fuel to common rail 64 . The fuel pump 65 is a plunger pump driven by the crankshaft 15 in the engine 1 of this configuration example. Common rail 64 stores fuel sent from fuel pump 65 at high fuel pressure. When the injector 6 opens, the fuel stored in the common rail 64 is injected into the combustion chamber 17 from the injection port of the injector 6 . The fuel supply system 61 can supply high pressure fuel of 30 MPa or higher to the injector 6 . The pressure of fuel supplied to the injector 6 may be changed according to the operating state of the engine 1 . Note that the configuration of the fuel supply system 61 is not limited to the configuration described above.

A spark plug 25 is attached to the cylinder head 13 for each cylinder 11 . A spark plug 25 forcibly ignites the air-fuel mixture in the combustion chamber 17 . Although not shown in detail, the electrode of the spark plug 25 faces the inside of the combustion chamber 17 and is positioned near the ceiling surface of the combustion chamber 17 .

An intake passage 40 is connected to one side surface of the engine 1 . The intake passage 40 communicates with the intake port 18 of each cylinder 11 . Gas introduced into the combustion chamber 17 flows through the intake passage 40 . An air cleaner 41 is arranged at the upstream end of the intake passage 40 . The air cleaner 41 filters fresh air. A surge tank 42 is arranged near the downstream end of the intake passage 40 . The intake passage 40 downstream of the surge tank 42 constitutes an independent passage that branches for each cylinder 11 . A downstream end of the independent passage is connected to the intake port 18 of each cylinder 11 .

A throttle valve 43 is arranged between the air cleaner 41 and the surge tank 42 in the intake passage 40 . The throttle valve 43 adjusts the amount of fresh air introduced into the combustion chamber 17 by adjusting the opening of the valve.

A supercharger 44 is also arranged in the intake passage 40 downstream of the throttle valve 43 . The supercharger 44 supercharges the gas introduced into the combustion chamber 17 . In the engine 1 of this configuration example, the supercharger 44 is a mechanical supercharger driven by the engine 1 . The mechanical supercharger 44 may be of the Roots, Lysholm, vane, or centrifugal type.

An electromagnetic clutch 45 is interposed between the supercharger 44 and the engine 1 . Between the supercharger 44 and the engine 1 , the electromagnetic clutch 45 transmits driving force from the engine 1 to the supercharger 44 or cuts off transmission of the driving force. The supercharger 44 is switched between on and off by the ECU 10 switching between connection and non-connection of the electromagnetic clutch 45 .

An intercooler 46 is arranged downstream of the supercharger 44 in the intake passage 40 . The intercooler 46 cools the gas compressed by the supercharger 44 . The intercooler 46 may be configured to be water-cooled or oil-cooled, for example.

A bypass passage 47 is connected to the intake passage 40 . The bypass passage 47 connects the upstream portion of the supercharger 44 and the downstream portion of the intercooler 46 in the intake passage 40 so as to bypass the supercharger 44 and the intercooler 46 . An air bypass valve 48 is arranged in the bypass passage 47 . The air bypass valve 48 adjusts the flow rate of gas flowing through the bypass passage 47 .

The ECU 10 fully opens the air bypass valve 48 when the supercharger 44 is turned off (that is, when the electromagnetic clutch 45 is disconnected). Gas flowing through the intake passage 40 bypasses the supercharger 44 and is introduced into the combustion chamber 17 of the engine 1 . The engine 1 operates in a non-supercharged state, that is, in a naturally aspirated state.

When the supercharger 44 is turned on, the engine 1 operates in a supercharged state. The ECU 10 adjusts the opening degree of the air bypass valve 48 when the supercharger 44 is turned on (that is, when the electromagnetic clutch 45 is connected). Part of the gas that has passed through the supercharger 44 flows back upstream of the supercharger 44 through the bypass passage 47 . When the ECU 10 adjusts the opening of the air bypass valve 48, the boost pressure of the gas introduced into the combustion chamber 17 changes. It should be noted that the term "supercharging" refers to the time when the pressure in the surge tank 42 exceeds the atmospheric pressure, and the term "non-supercharging" refers to the time when the pressure in the surge tank 42 falls below the atmospheric pressure. good too.

In the engine 1 of this configuration example, a supercharging system 49 is configured by the supercharger 44 , the bypass passage 47 and the air bypass valve 48 . The supercharger 44 is turned off, for example, in a low-load and low-speed operating region, and turned on in other operating regions.

The engine 1 has a swirl generator in the combustion chamber 17 that generates a swirl flow. The swirl generator has a swirl control valve 56 attached to the intake passage 40 . Although detailed illustration is omitted, the swirl control valve 56 is disposed in the secondary passage of the primary passage leading to the first intake port of the two intake ports 18 and the secondary passage leading to the second intake port. It is The swirl control valve 56 is an opening control valve capable of narrowing the cross section of the secondary passage. When the opening of the swirl control valve 56 is small, the amount of intake air entering the combustion chamber 17 from the first intake port is relatively large and the amount of intake air entering the combustion chamber 17 from the second intake port is relatively small. The swirl flow in chamber 17 becomes stronger. When the opening of the swirl control valve 56 is large, the amount of intake air entering the combustion chamber 17 from the first intake port and the second intake port becomes substantially equal, so the swirl flow in the combustion chamber 17 becomes weak. When the swirl control valve 56 is fully opened, no swirl flow occurs.

An exhaust passage 50 is connected to the other side surface of the engine 1 . The exhaust passage 50 communicates with the exhaust port 19 of each cylinder 11 . The exhaust passage 50 is a passage through which the exhaust gas discharged from the combustion chamber 17 flows. An upstream portion of the exhaust passage 50 constitutes an independent passage that branches for each cylinder 11, although detailed illustration is omitted. The upstream end of the independent passage is connected to the exhaust port 19 of each cylinder 11 .

An exhaust gas aftertreatment system with a plurality of catalytic converters is arranged in the exhaust passage 50 . Although not shown, the upstream catalytic converter is arranged in the engine room. The upstream catalytic converter has a three-way catalyst 511 and a GPF (Gasoline Particulate Filter) 512 . A downstream catalytic converter is also located in the engine compartment. The downstream catalytic converter has a three-way catalyst 513 . It should be noted that the exhaust gas aftertreatment system is not limited to the configuration of the illustrated example. For example, GPF may be omitted. Also, the catalytic converter is not limited to having a three-way catalyst. Furthermore, the order in which the three-way catalyst and GPF are arranged may be changed as appropriate.

An EGR passage 52 that constitutes an external EGR system is connected between the intake passage 40 and the exhaust passage 50 . The EGR passage 52 is a passage for recirculating part of the exhaust gas to the intake passage 40 . The upstream end of the EGR passage 52 is connected to a portion downstream of the downstream catalytic converter in the exhaust passage 50 . A downstream end of the EGR passage 52 is connected to an upstream portion of the supercharger 44 in the intake passage 40 .

A water-cooled EGR cooler 53 is arranged in the EGR passage 52 . The EGR cooler 53 cools the exhaust gas. An EGR valve 54 is also arranged in the EGR passage 52 . The EGR valve 54 adjusts the flow rate of exhaust gas flowing through the EGR passage 52 . By adjusting the opening of the EGR valve 54, the recirculation amount of the cooled exhaust gas, that is, the external EGR gas can be adjusted.

In the engine 1 of this configuration example, the EGR system 55 is composed of an external EGR system and an internal EGR system. The external EGR system can supply cooler exhaust gas to the combustion chamber 17 than the internal EGR system.

The engine 1 has a cooling system 70 . The cooling system 70 includes a pump 71 that supplies coolant, an inlet passage 72 that flows the coolant from the pump 71 into the cylinder block 12 of the engine 1 , and a coolant passage in the engine 1 . A radiator passage 74 through which the coolant flows from the head 13 through the radiator 73 to the pump 71, and a radiator bypass through which the coolant flows into the pump 71 after passing through the coolant passage in the engine 1, bypassing the radiator 73. passageway 75.

The pump 71 is a mechanical pump driven in conjunction with the crankshaft 15 . A discharge port of the pump 71 is connected to the inlet passage 72 . The pump 71 is provided with a first liquid temperature sensor SW9 that detects the liquid temperature of the cooling liquid discharged to the inlet passage 72 . The amount of coolant discharged from the pump 71 varies depending on the engine speed and the amount of coolant returned to the pump 71 . The first liquid temperature sensor SW9 may be arranged so as to detect the liquid temperature of the cooling liquid flowing through the inlet passage 72 .

The inlet passage 72 communicates the outlet of the pump 71 and the inlet of the coolant passage of the cylinder block 12 .

The radiator passage 74 is connected to the coolant passage of the cylinder head 13 . A thermostat valve 76 is arranged between the radiator 73 and the pump 71 in the radiator passage 74 . The thermostat valve 76 is composed of an electric thermostat valve. Specifically, the thermostat valve 76 is a general thermostat valve with a built-in heating wire. The thermostat valve 76 is configured to open according to the temperature of the cooling liquid when the temperature of the cooling liquid is equal to or higher than a predetermined liquid temperature when no power is supplied. can be opened even when the temperature of the liquid is lower than the predetermined liquid temperature. That is, when the power is off, the thermostat valve 76 opens at a predetermined liquid temperature, so that the temperature of the coolant in the radiator passage 74 can be brought to around the predetermined liquid temperature. By opening the thermostat valve 76 at the liquid temperature, the liquid temperature of the cooling liquid in the radiator passage 74 can be adjusted to the desired liquid temperature.

The radiator bypass passage 75 is also connected to the cooling liquid passage of the cylinder head 13 in the same manner as the radiator passage 74 . A flow control valve 77 is arranged in the middle of the radiator bypass passage 75 . The flow control valve 77 is an on/off valve that can be switched between an open state with a constant degree of opening and a closed state with a fully closed state. The flow control valve 77 adjusts the time in the open state and the time in the closed state, more specifically, by adjusting the ratio of the open state and the closed state per unit time (hereinafter referred to as duty ratio). The flow rate of the cooling liquid passing through the bypass passage 75 is adjusted.

(Configuration of engine control device)
The engine control device includes an ECU (Engine Control Unit) 10 . The ECU 10 is a well-known microcomputer-based controller, and as shown in FIG. (Read Only Memory) for storing programs and data, and an input/output bus 103 for inputting/outputting electric signals. The ECU 10 is an example of a control section.

Various sensors SW1 to SW17 are connected to the ECU 10 as shown in FIGS. The sensors SW1 to SW17 output signals to the ECU 10. FIG. The sensors include the following sensors.

Airflow sensor SW1: arranged downstream of the air cleaner 41 in the intake passage 40, and measures the flow rate of fresh air flowing through the intake passage 40. First intake air temperature sensor SW2: arranged downstream of the air cleaner 41 in the intake passage 40, and , to measure the temperature of fresh air flowing through the intake passage 40. First pressure sensor SW3: arranged downstream of the connecting position of the EGR passage 52 in the intake passage 40 and upstream of the supercharger 44, and Second intake air temperature sensor SW4: arranged downstream of the turbocharger 44 in the intake passage 40 and upstream of the connecting position of the bypass passage 47, and flows out of the turbocharger 44 Second pressure sensor SW5 for measuring gas temperature Second pressure sensor SW5: Attached to surge tank 42 for measuring pressure of gas downstream of supercharger 44 In-cylinder pressure sensor SW6: Attached to cylinder head 13 corresponding to each cylinder 11 and measures the pressure in each combustion chamber 17 Linear _O2 sensor SW7: Arranged upstream of the upstream catalytic converter in the exhaust passage 50 and measures the oxygen concentration in the exhaust gas Lambda _O2 sensor SW8: Arranged downstream of the three-way catalyst 511 in the upstream catalytic converter and measures the oxygen concentration in the exhaust gas First liquid temperature sensor SW9: As described above, it is attached to the pump 71 and flows into the cylinder block 12 Second liquid temperature sensor SW10: mounted on the engine 1 to measure the temperature of the coolant immediately after heat exchange with the engine 1. Crank angle sensor SW11: mounted on the engine 1 and detects the temperature of the coolant. Measures the rotation angle of the shaft 15 Accelerator opening sensor SW12: Attached to the accelerator pedal mechanism and measures the accelerator opening corresponding to the amount of operation of the accelerator pedal Intake cam angle sensor SW13: Attached to the engine 1 and used for intake Measuring the rotation angle of the camshaft Exhaust cam angle sensor SW14: Attached to the engine 1 and measuring the rotation angle of the exhaust camshaft EGR differential pressure sensor SW15: Arranged in the EGR passage 52 and upstream and downstream of the EGR valve 54 A fuel pressure sensor SW16: mounted on the common rail 64 of the fuel supply system 61 to measure the pressure of the fuel supplied to the injectors 6. A third intake air temperature sensor SW17: mounted on the surge tank 42 and The temperature of the gas in the tank 42, in other words, the temperature of the intake air introduced into the combustion chamber 17 is measured.

The ECU 10 determines the operating state of the engine 1 based on the signals from these sensors SW1 to SW17, and calculates control amounts for each device according to predetermined control logic. The control logic is stored in memory 102 . The control logic includes using the operational map stored in memory 102 to compute target and/or control variables.

The ECU 100 transmits electric signals related to the calculated control amount to the injector 6, the spark plug 25, the electric intake S-VT 23, the electric exhaust S-VT 24, the fuel supply system 61, the throttle valve 43, the EGR valve 54, the supercharger 44. , the electromagnetic clutch 45 , the air bypass valve 48 , the swirl control valve 56 , the thermostat valve 76 and the flow control valve 77 .

For example, the ECU 10 sets the target torque of the engine 1 and determines the target supercharging pressure based on the signal from the accelerator opening sensor SW12 and the driving map. Then, the ECU 10 adjusts the opening degree of the air bypass valve 48 based on the target supercharging pressure and the differential pressure across the supercharger 44 obtained from the signals of the first pressure sensor SW3 and the second pressure sensor SW5. By performing feedback control, the supercharging pressure is brought to the target supercharging pressure.

Further, the ECU 10 sets the target EGR rate based on the operating state of the engine 1 and the operating map. EGR rate is the ratio of EGR gas to total gas in combustion chamber 17 . The ECU 10 determines the target EGR gas amount based on the target EGR rate and the intake air amount based on the signal of the accelerator opening sensor SW12, and based on the differential pressure across the EGR valve 54 obtained from the signal of the EGR differential pressure sensor SW15. The external EGR gas amount introduced into the combustion chamber 17 is made to match the target EGR gas amount by performing feedback control for adjusting the opening degree of the EGR valve 54 by means of the external EGR gas amount.

Furthermore, the ECU 10 executes air-fuel ratio feedback control when a predetermined control condition is satisfied. Specifically, the ECU 10 controls the fuel injection of the injector 6 so that the air-fuel ratio of the air-fuel mixture becomes a desired value based on the oxygen concentration in the exhaust gas measured by the linear _O2 sensor SW7 and the lambda _O2 sensor SW8. Adjust quantity.

(SPCCI combustion concept)
The main purpose of the engine 1 is to improve fuel efficiency and exhaust gas performance. In the self-ignition combustion, if the temperature in the combustion chamber 17 before the start of compression varies, the self-ignition timing changes greatly. Therefore, the engine 1 performs SPCCI combustion, which is a combination of SI combustion and CI combustion.

In the SPCCI combustion, the ignition plug 25 forcibly ignites the air-fuel mixture in the combustion chamber 17, and the air-fuel mixture undergoes SI combustion due to flame propagation, and the heat generated by the SI combustion causes the combustion in the combustion chamber 17. When the temperature rises and the pressure in the combustion chamber 17 rises due to flame propagation, the unburned air-fuel mixture undergoes self-ignition for CI combustion.

By adjusting the calorific value of SI combustion, it is possible to absorb temperature variations in the combustion chamber 17 before the start of compression. By adjusting the ignition timing by the ECU 10, the air-fuel mixture can be self-ignited at the target timing.

In SPCCI combustion, the heat release during SI combustion is milder than that during CI combustion. FIG. 3 illustrates a heat release rate waveform 301 in SPCCI combustion. The heat release rate waveform 301 in SPCCI combustion has a smaller rise slope than the rise slope in the CI combustion waveform. Further, the pressure fluctuation (dp/dθ) in the combustion chamber 17 is also gentler during SI combustion than during CI combustion.

When the unburned air-fuel mixture self-ignites after the start of SI combustion, the slope of the heat release rate waveform may change from small to large at the timing of self-ignition. The heat release rate waveform may have an inflection point X at the timing θci at which CI combustion starts.

After the start of CI combustion, SI combustion and CI combustion are performed in parallel. Since CI combustion produces more heat than SI combustion, the heat release rate is relatively high. However, since the CI combustion is performed after the top dead center of the compression stroke, the slope of the heat release rate waveform is prevented from becoming too large. The pressure fluctuation (dp/dθ) during CI combustion is also relatively mild.

Pressure fluctuation (dp/dθ) can be used as an index representing combustion noise. As described above, the SPCCI combustion has a small pressure fluctuation (dp/dθ), so excessive combustion noise can be avoided. Combustion noise of the engine 1 is suppressed below the permissible level.

The SPCCI combustion ends when the CI combustion ends. CI combustion has a shorter combustion period than SI combustion alone, illustrated by waveform 302 in FIG. The SPCCI combustion has an earlier combustion end timing than the SI combustion.

(engine operating range)
FIG. 4 illustrates an operation map 401 related to control of the engine 1. As shown in FIG. The driving map 401 is stored in the memory 102 of the ECU 10. FIG. The operation map 401 is an operation map when the engine 1 is warm.

The operation map 401 is defined by the load and rotation speed of the engine 1 . The operation map 401 is divided into three regions depending on the rotation speed and the load, as indicated by solid lines in FIG. 4 . Specifically, the operation map 401 includes a low-load area A1 in the low- and medium-rotation areas, medium- and high-load areas A2, A3, and A4 in the low- and medium-rotation areas, and at the rotation speed N1 or higher. It is divided into a certain high rotation area A5.

Here, the low speed region, the middle speed region, and the high speed region are obtained by dividing the entire operating region of the engine 1 in the rotational speed direction into approximately three equal parts: the low speed region, the middle speed region, and the high speed region. , a low rotation area, a middle rotation area, and a high rotation area. In addition, the low load range, the medium load range, and the high load range each divide the entire operating range of the engine 1 in the load direction into approximately three equal parts: the low load range, the medium load range, and the high load range. A low load range, a medium load range, and a high load range may also be used.

The operation map 401 indicates the state of the air-fuel mixture and the combustion mode in each region. The engine 1 performs SPCCI combustion in the regions A1, A2, A3, and A4. Engine 1 also performs SI combustion in region A5. Hereinafter, the operation of the engine 1 in each region of the operation map 401 of FIG. 4 will be described in detail with reference to the fuel injection timing and ignition timing shown in FIG. The horizontal axis of FIG. 5 is the crank angle. Reference numerals 601, 602, 603, 604, 605 and 606 in FIG. 5 respectively correspond to the operating states of the engine 1 indicated by reference numerals 601, 602, 603, 604, 605 and 606 in the operation map 401 of FIG.

(Engine operation in area A1)
When the engine 1 is operating in the area A1, the engine 1 performs SPCCI combustion. Reference numeral 601 in FIG. 5 indicates the fuel injection timing (reference numerals 6011 and 6012) and the ignition timing (reference numeral 6013) when the engine 1 is operating in the operating state 601 in the region A1, and the combustion waveform (that is, the crank angle 6014) showing the change in heat release rate with respect to . Reference numeral 602 denotes fuel injection timing (reference numerals 6021 and 6022), ignition timing (reference numeral 6023), and combustion waveform (reference numeral 6024) when the engine 1 is operating in the operating state 602 in the region A1. Reference numeral 603 indicates the fuel injection timings (reference numerals 6031 and 6032), the ignition timing (reference numeral 6033), and the combustion waveform (reference numeral 6034) when the engine 1 is operating in the operating state 603 in the region A1. The operating states 601, 602, and 603 have the same rotational speed of the engine 1 and different loads. The operating state 601 has the lowest load (ie, light load), the operating state 602 has the next lowest load (ie, low load), and the operating state 603 has the highest load among them.

The EGR system 55 introduces EGR gas into the combustion chamber 17 in order to improve fuel efficiency of the engine 1 . Specifically, the electric intake S-VT 23 and the electric exhaust S-VT 24 provide a positive overlap period in which both the intake valve 21 and the exhaust valve 22 are opened near exhaust top dead center. A portion of the exhaust gas discharged from the combustion chamber 17 to the intake port 18 and the exhaust port 19 is reintroduced into the combustion chamber 17 . Due to the introduction of hot exhaust gases into the combustion chamber 17, the temperature in the combustion chamber 17 increases. This is advantageous for stabilizing SPCCI combustion. Incidentally, the electric intake S-VT 23 and the electric exhaust S-VT 24 may be provided with a negative overlap period in which both the intake valve 21 and the exhaust valve 22 are closed.

Also, the swirl generator forms a strong swirl flow in the combustion chamber 17 . The swirl ratio is, for example, 4 or more. The swirl control valve 56 is fully closed or has a predetermined degree of opening on the closing side. As described above, since the intake port 18 is a tumble port, an oblique swirl flow having a tumble component and a swirl component is formed in the combustion chamber 17 .

The injector 6 injects fuel into the combustion chamber 17 several times (6011, 6012, 6021, 6022, 6031, 6032) during the intake stroke. The multiple fuel injections and the swirl flow in the combustion chamber 17 stratify the air-fuel mixture.

The fuel concentration of the air-fuel mixture in the central portion of the combustion chamber 17 is higher than the fuel concentration in the outer peripheral portion. Specifically, the air-fuel ratio (A/F) of the air-fuel mixture in the central portion is 20 or more and 30 or less, and the air-fuel ratio (A/F) of the air-fuel mixture in the outer peripheral portion is 35 or more. The value of the air-fuel ratio is the value of the air-fuel ratio at the time of ignition, and the same applies to the following description. By setting the A/F of the air-fuel mixture close to the spark plug 25 to 20 or more and 30 or less, it is possible to suppress the generation of RawNOx during SI combustion. Further, CI combustion is stabilized by setting the A/F of the air-fuel mixture at the outer periphery to 35 or higher.

The A/F of the air-fuel mixture is leaner than the stoichiometric air-fuel ratio in the entire combustion chamber 17 (that is, excess air ratio λ>1). More specifically, the A/F of the air-fuel mixture in the entire combustion chamber 17 is 30 or higher. By doing so, the generation of RawNOx can be suppressed, and the exhaust gas performance can be improved.

When the load of the engine 1 is low (that is, in the operating state 601), the injector 6 performs the first injection 6011 in the first half of the intake stroke and the second injection 6012 in the second half of the intake stroke. The first half of the intake stroke may be the first half when the intake stroke is divided into the first half and the second half, and the second half of the intake stroke may be the second half when the intake stroke is divided into two halves. Also, the injection amount ratio between the first injection 6011 and the second injection 6012 may be set to 9:1, for example.

In the operating state 602 in which the load of the engine 1 is high, the injector 6 starts the second injection 6022 performed in the second half of the intake stroke at a timing advanced from the second injection 6012 in the operating state 601 . By advancing the second injection 6022, the air-fuel mixture in the combustion chamber 17 approaches homogeneity. The injection quantity ratio between the first injection 6021 and the second injection 6022 may be, for example, 7:3 to 8:2.

In the operating state 603 in which the load of the engine 1 is even higher, the injector 6 starts the second injection 6032 performed in the second half of the intake stroke at a timing further advanced than the second injection 6022 in the operating state 602 . By further advancing the second injection 6032, the air-fuel mixture in the combustion chamber 17 becomes even closer to homogeneity. The injection amount ratio between the first injection 6031 and the second injection 6032 may be 6:4, for example.

After the end of fuel injection, the ignition plug 25 ignites the air-fuel mixture in the central portion of the combustion chamber 17 at a predetermined timing before compression top dead center (reference numerals 6013, 6023, 6033). The ignition timing may be the end of the compression stroke. The final stage of the compression stroke may be the final stage when the compression stroke is divided into three equal parts, the initial stage, the middle stage, and the final stage.

As described above, since the air-fuel mixture in the central portion has a relatively high fuel concentration, the ignitability is improved and the SI combustion due to flame propagation is stabilized. By stabilizing SI combustion, CI combustion starts at an appropriate timing. Controllability of CI combustion is improved. Generation of combustion noise is suppressed. Further, by making the A/F of the air-fuel mixture lean and performing SPCCI combustion, the fuel efficiency of the engine 1 can be greatly improved.

(Engine operation in areas A2, A3, and A4)
When the engine 1 is operating in the regions A2, A3 and A4, the engine 1 performs SPCCI combustion. Reference numeral 604 in FIG. 5 indicates the fuel injection timing (reference numerals 6041 and 6042), the ignition timing (reference numeral 6043), and the combustion waveform (reference numeral 6044) when the engine 1 is operating in the operating state 604 in the region A2. showing. Reference numeral 605 indicates the fuel injection timing (reference numeral 6051), the ignition timing (reference numeral 6052), and the combustion waveform (reference numeral 6053) when the engine 1 is operating in the operating state 605 in the region A4.

EGR system 55 introduces EGR gases into combustion chamber 17 . Specifically, the electric intake S-VT 23 and the electric exhaust S-VT 24 provide a positive overlap period in which both the intake valve 21 and the exhaust valve 22 are opened near exhaust top dead center. Internal EGR gases are introduced into the combustion chamber 17 . Also, the EGR system 55 introduces the exhaust gas cooled by the EGR cooler 53 into the combustion chamber 17 through the EGR passage 52 . That is, the external EGR gas having a lower temperature than the internal EGR gas is introduced into the combustion chamber 17 . External EGR gas regulates the temperature in combustion chamber 17 to a suitable temperature. The EGR system 55 reduces the amount of EGR gas as the load on the engine 1 increases. The EGR system 55 may zero EGR gas, including internal EGR gas and external EGR gas, at full load.

Also, in the regions A2 and A3, the swirl control valve 56 is fully closed or at a predetermined degree of opening on the closed side. A strong swirl flow with a swirl ratio of 4 or more is formed in the combustion chamber 17 . On the other hand, in region A4, swirl control valve 56 is open.

The A/F of the air-fuel mixture in the entire combustion chamber 17 is the stoichiometric air-fuel ratio (A/F≈14.7). By cleaning the exhaust gas discharged from the combustion chamber 17 by the three-way catalysts 511 and 513, the exhaust gas performance of the engine 1 is improved. The A/F of the air-fuel mixture should be within the cleaning window of the three-way catalyst. The excess air ratio λ of the mixture may be 1.0±0.2. Incidentally, when the engine 1 is operating in the region A3 including the maximum load, the A/F of the air-fuel mixture in the entire combustion chamber 17 may be the stoichiometric air-fuel ratio or richer than the stoichiometric air-fuel ratio (that is, The excess air ratio λ of the air-fuel mixture is λ≦1).

Since the EGR gas is introduced into the combustion chamber 17, G/F, which is the weight ratio of the total gas and fuel in the combustion chamber 17, becomes leaner than the stoichiometric air-fuel ratio. The G/F of the air-fuel mixture may be 18 or more. By doing so, it is possible to avoid the occurrence of so-called knocking. G/F may be set to 18 or more and 30 or less. Also, G/F may be set to 18 or more and 50 or less.

When the engine 1 operates in the operating state 604, the injector 6 performs multiple fuel injections (reference numerals 6041, 6042) during the intake stroke. The injector 6 may perform the first injection 6041 in the first half of the intake stroke and perform the second injection 6042 in the second half of the intake stroke.

Further, when the engine 1 operates in the operating state 605, the injector 6 injects fuel during the intake stroke (reference numeral 6051).

The spark plug 25 ignites the air-fuel mixture at a predetermined timing near the compression top dead center after fuel injection (reference numerals 6043 and 6052). When the engine 1 is running in the operating state 604, the spark plug 25 may ignite before compression top dead center (reference numeral 6043). When the engine 1 is running in the operating state 605, the spark plug 25 may ignite after compression top dead center (reference numeral 6052).

By performing SPCCI combustion with the A/F of the air-fuel mixture at the stoichiometric air-fuel ratio, the three-way catalysts 511 and 513 can be used to clean the exhaust gas discharged from the combustion chamber 17 . Further, by introducing the EGR gas into the combustion chamber 17 to dilute the air-fuel mixture, the fuel efficiency of the engine 1 is improved.

(Engine operation in area A5)
When the rotation speed of the engine 1 is high, the time required for the crank angle to change by 1° is shortened. When the rotation speed of the engine 1 increases, it becomes difficult to stratify the air-fuel mixture in the combustion chamber 17, making it difficult to perform SPCCI combustion. Therefore, when the engine 1 is operating in the area A5, the engine 1 performs SI combustion instead of SPCCI combustion.

Reference numeral 606 in FIG. 5 indicates the fuel injection timing (reference numeral 6061), the ignition timing (reference numeral 6062), and the combustion waveform (reference numeral 6063) when the engine 1 is operating in the high-load operating state 606 in the region A5. is shown.

EGR system 55 introduces EGR gases into combustion chamber 17 . EGR system 55 reduces the amount of EGR gas as the load increases. The EGR system 55 may zero EGR gas at full load.

Swirl control valve 56 is fully open. A swirl flow is not generated in the combustion chamber 17, and only a tumble flow is generated. By fully opening the swirl control valve 56, it is possible to improve charging efficiency and reduce pump loss.

The A/F of the air-fuel mixture is basically the stoichiometric air-fuel ratio (A/F≈14.7) in the entire combustion chamber 17 . The excess air ratio λ of the air-fuel mixture may be set to 1.0±0.2. The excess air ratio λ of the air-fuel mixture may be less than 1 when the engine 1 is operating near a fully open load.

The injector 6 starts fuel injection during the intake stroke. The injector 6 injects fuel all at once (reference numeral 6061). A homogeneous or substantially homogeneous mixture is formed in the combustion chamber 17 by initiating fuel injection during the intake stroke. In addition, since it is possible to secure a long vaporization time of the fuel, it is possible to reduce the unburned loss.

The spark plug 25 ignites the air-fuel mixture at an appropriate timing after the end of fuel injection and before the top dead center of the compression stroke (reference numeral 6062).

(Estimation of Exhaust Gas Temperature)
Regarding the control of the engine 1, the temperature of the exhaust gas is used for various purposes. For example, if the temperature of the catalyst of the exhaust gas aftertreatment system is too high, the reliability of the catalyst is reduced. The temperature of the exhaust gas entering the catalyst must be monitored and the temperature of the exhaust gas must be reduced if the temperature of the exhaust gas is too high. Further, for example, in order to adjust the recirculated amount of EGR gas, the ECU 10 must grasp the temperature of the exhaust gas.

Attaching a temperature sensor to the exhaust passage 50 of the engine 1 increases the cost. If multiple sensors are used to measure the temperature of the exhaust gas entering the catalyst and the temperature of the exhaust gas entering the EGR passage 52, the cost is further increased. Therefore, the control device of the engine 1 is configured to estimate the temperature of the exhaust gas at the most upstream portion of the exhaust passage 50 based on the combustion state of the air-fuel mixture in the cylinder without attaching a temperature sensor to the exhaust passage 50. It is By estimating the temperature of the exhaust gas at the most upstream portion of the exhaust passage 50, the control device determines the temperature of the exhaust gas flowing into the catalyst and the exhaust gas flowing into the EGR passage 52 based on the temperature of the exhaust gas at the most upstream portion. The temperature of the gas, etc. can also be estimated.

FIG. 6 illustrates functional blocks of the ECU 10 related to estimation of exhaust gas temperature. The ECU 10 has an acquisition section 104 , a calculation section 105 , an estimation section 106 and a correction section 107 .

Acquisition unit 104 acquires a signal from in-cylinder pressure sensor SW6 and a signal from crank angle sensor SW11.

Based on the signal from the in-cylinder pressure sensor SW6 and the signal from the crank angle sensor SW11 acquired by the acquisition unit 104, the calculation unit 105 calculates the degree of combustion progress, which is the crank angle when the combustion in the cylinder progresses to a specific degree. calculate. The calculator 105 may specifically calculate the crank angle at which the mass combustion ratio is 50%, that is, mfb50 (or the combustion center of gravity). A known method may be used as a method for calculating mfb50. Specifically, the calculation unit 105 can calculate mfb50 from the area of the heat release rate waveform 301 or 302 obtained from the signal of the in-cylinder pressure sensor SW6 and the signal of the crank angle sensor SW11, as illustrated in FIG. can. If there is a correlation with the exhaust gas temperature, the value of the mass combustion ratio that can be used as the degree of combustion progress can be any value such as mfb10 or mfb90.

Note that the calculation unit 105 may calculate the combustion center of gravity as the degree of combustion progress based on the signal from the crank angle sensor SW11, for example. For example, waveform 303 in FIG. 3 exemplifies the time change of the signal of crank angle sensor SW11. When CI combustion starts during SPCCI combustion, the crank speed increases. In addition, in the waveform 301 of the heat release rate of SPCCI combustion, the inflection point X and mfb50 are crank angles close to each other. Therefore, mfb50 may be calculated based on the pulse interval of the crank angle sensor SW11.

The estimation unit 106 estimates the temperature of the exhaust gas based on the relationship between the degree of progress of combustion and the temperature of the exhaust gas. More specifically, the estimation unit 106 estimates the temperature of the exhaust gas from the degree of combustion progress calculated by the calculation unit 105, using a model representing the relationship between the degree of combustion progress and the temperature of the exhaust gas. The model is stored in memory 102 . The model is specifically configured as shown in FIGS. 8, 9 and 11. FIG. The details of the configuration of the model will be described later.

The correction unit 107 corrects the exhaust gas temperature estimated by the estimation unit 106 according to the temperature of the engine 1 . Details of correction by the correction unit 107 will be described later.

Next, a procedure for estimating the temperature of the exhaust gas executed by the ECU 10 will be described with reference to the flowchart of FIG. Note that the order of each step in the flowchart of FIG. 7 can be changed.

In step S1, the ECU 10 reads signals from the sensors SW1 to SW17, and in subsequent step S2, the ECU 10 executes fuel injection and ignition corresponding to the operating state of the engine 1 (see FIGS. 4 and 5).

Next, in step S3, the acquisition unit 104 of the ECU 10 acquires signals from the in-cylinder pressure sensor SW6 and the crank angle sensor SW11 during in-cylinder combustion, and in subsequent step S4, the calculation unit 105 calculates the degree of combustion progress. .

In step S5, the estimation unit 106 of the ECU 10 determines whether or not the combustion mode of the engine 1 is SI combustion. If the combustion mode is SI combustion and the determination in step S5 is YES, the process proceeds to step S6, and if the combustion mode is SPCCI combustion and the determination in step S5 is NO, the process proceeds to step S8.

In step S8, the estimation unit 106 determines whether the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio. If the air-fuel ratio is the stoichiometric air-fuel ratio (ie λ=1) and the determination in step S8 is YES, the process proceeds to step S9 where the air-fuel ratio is lean (ie λ>1) and the determination in step S8 is If NO, the process proceeds to step S11.

In step S6, the estimation unit 106 estimates the temperature of the exhaust gas from the model stored in the memory 102 and the calculated degree of combustion progress. FIG. 8 illustrates a model representing the relationship between the degree of combustion progress and the exhaust gas temperature. FIG. 8 illustrates a model 801 at the time of SI combustion, a model 802 at the time of SPCCI combustion and the air-fuel ratio is the stoichiometric air-fuel ratio, and a model 803 at the time of SPCCI combustion and the air-fuel ratio is lean. It is

As can be seen from FIG. 8, the model is represented by a linear function. This is because the inventors of the present application have found that there is a linear correlation between the degree of combustion progress and the exhaust gas temperature when the degree of combustion progress is retarded relative to the degree of combustion progress in the MBT. Depends on what you find.

SI and SPCCI combustion ignite the mixture. When the ignition timing is retarded, the temperature of the exhaust gas rises. A correlation also exists between ignition timing and exhaust gas temperature. However, according to the study by the inventors of the present application, in a so-called spark ignition engine that performs SI combustion and/or SPCCI combustion, the relationship between the ignition timing and the exhaust gas temperature is not linear but nonlinear.

If the ignition timing is late, combustion starts in the expansion stroke. The expansion stroke is a stroke in which the volume inside the cylinder is expanded. Volume expansion has a non-linear relationship to crank angle progression. Combustion in the expansion stroke is subject to non-linear volumetric expansion. The degree of combustion progress has a nonlinear relationship (for example, a quadratic function) with respect to the ignition timing delay.

Since the degree of combustion progress is a parameter representing the state of combustion, there is a linear correlation between the degree of combustion progress and the indicated work of the engine 1 . Also, the temperature of the exhaust gas is determined by the heat amount obtained by subtracting the indicated work of the engine from the heat amount generated by combustion in the cylinder. Therefore, a linear correlation is also established between the indicated work of the engine and the temperature of the exhaust gas. Since the ignition timing and the degree of combustion progress have a non-linear relationship, the correlation between the ignition timing and the temperature of the exhaust gas is non-linear, but the correlation between the degree of combustion progress and the temperature of the exhaust gas is linear. Become.

Therefore, each of the models 801, 802, and 803 linearly estimates the exhaust gas temperature higher as the combustion progress is retarded than the MBT combustion progress. is configured to According to models 801, 802, and 803, estimating unit 106 linearly estimates the temperature of the exhaust gas to be higher as the degree of combustion retards more than the degree of combustion in MBT when the degree of combustion retards. do.

Also, if the ignition timing is advanced too much, abnormal combustion may occur. A knock limit is set as an ignition timing advance limit at which abnormal combustion does not occur. The knock limit is set on the retard side of MBT. The ignition timing is set to retard the knock limit.

As shown in FIG. 8, even when the degree of combustion progresses more retarded than the degree of combustion progress at the knock limit, a linear correlation is established between the degree of combustion progress and the exhaust gas temperature. Models 801, 802, and 803 each linearly estimate the exhaust gas temperature higher as the combustion progress is retarded than the combustion progress at the knock limit. It is configured. In accordance with models 801, 802, and 803, estimating unit 106 linearly increases the exhaust gas temperature as the combustion progress is retarded relative to the combustion progress at the knock limit. presume.

Since the models 801, 802, and 803 are linear, the estimator 106 can accurately estimate the temperature of the exhaust gas even when the degree of combustion lags significantly behind the degree of combustion in the MBT.

The model 801 for SI combustion and the models 802 and 803 for SPCCI combustion are different. A model 801 represents a first relationship between the combustion progress and exhaust gas temperature during SI combustion. Models 802, 803 represent a second relationship between combustion progress and exhaust gas temperature during SPCCI combustion. Estimating section 106 estimates the temperature of the exhaust gas by switching models 801, 802, and 803 corresponding to the switching of the combustion mode. As a result, the estimating unit 106 can accurately estimate the temperature of the exhaust gas during each of the SI combustion and the SPCCI combustion.

More specifically, the model 801 for SI combustion is configured to estimate the exhaust gas temperature higher than the models 802 and 803 for SPCCI combustion for the same degree of combustion. As a result, estimating section 106 estimates the exhaust gas temperature to be higher during SI combustion than during SPCCI combustion when the degree of combustion progress is the same.

SPCCI combustion burns part of the unburned air-fuel mixture by self-ignition, and therefore has higher thermal efficiency than SI combustion. Therefore, SPCCI combustion has a lower exhaust gas temperature than SI combustion, and SI combustion has a higher exhaust gas temperature than SPCCI combustion. The model 801 for SI combustion is configured to estimate the exhaust gas temperature higher than the models 802 and 803 for SPCCI combustion for the same degree of combustion progress. It is possible to accurately estimate the temperature of the exhaust gas at each time.

When the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio, the temperature rise rate of the exhaust gas in the model 802 during SPCCI combustion with respect to the change in the degree of combustion is higher than the temperature rise rate in the model 801 during SI combustion. is configured as "Rate of temperature rise of exhaust gas with respect to change in degree of combustion" is the slope of the straight line in FIG.

As described above, the thermal efficiency is high during SPCCI combustion. On the other hand, in SPCCI combustion, part of the air-fuel mixture undergoes CI combustion. CI combustion performed in the expansion stroke is greatly affected by changes in volume. Specifically, when the degree of combustion progress is retarded, the thermal efficiency of SPCCI combustion is greatly reduced. During SPCCI combustion, the temperature of the exhaust gas changes greatly with changes in the degree of combustion progress. In contrast, SI combustion due to flame propagation is less susceptible to volumetric changes. Even if the progress of combustion slows during SI combustion, the thermal efficiency of SI combustion does not significantly decrease. During SI combustion, the change in exhaust gas temperature is small with respect to the change in the degree of combustion progress.

Therefore, in the model 802 for SPCCI combustion, the temperature rise rate of the exhaust gas with respect to the change in the degree of combustion is made higher than in the model 801 for SI combustion. That is, model 802 has a steeper slope than model 801 . As a result, the estimation unit 106 can accurately estimate the temperature of the exhaust gas during SI combustion and SPCCI combustion.

Also, the models 801 and 802 when the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio differ from the model 803 when the air-fuel ratio is lean. Furthermore, during SPCCI combustion, the models 801 and 802 when the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio differ from the model 803 when the air-fuel ratio is lean. The estimating unit 106 can accurately estimate the temperature of the exhaust gas by switching between the models 802 and 803 corresponding to the switching of the air-fuel ratio of the air-fuel mixture during SPCCI combustion and estimating the temperature of the exhaust gas. .

More specifically, model 803 when the air/fuel mixture is lean is configured to estimate a lower exhaust gas temperature for the same degree of combustion than model 802 when the air/fuel ratio is stoichiometric. there is As a result, the estimating unit 106 estimates the temperature of the exhaust gas to be lower when the air-fuel ratio is lean than when the air-fuel ratio is the stoichiometric air-fuel ratio when the degree of combustion progress is the same.

When the air-fuel ratio of the air-fuel mixture is lean, the thermal efficiency of the engine 1 is relatively high, so the temperature of the exhaust gas is correspondingly lowered. The model 803 when lean is configured to estimate the exhaust gas temperature lower than the models 801 and 802 when the air-fuel ratio is stoichiometric for the same degree of combustion progress. The temperature of the exhaust gas can be accurately estimated when the air-fuel ratio is high and when the air-fuel ratio is lean.

The model 803 when the air-fuel ratio of the air-fuel mixture is lean is configured so that the temperature rise rate of the exhaust gas with respect to the change in the degree of combustion is lower than the temperature rise rate of the model 802 when the air-fuel ratio is stoichiometric. . That is, model 803 has a gentler slope than model 802 .

As described above, the thermal efficiency is high during SPCCI combustion, and the thermal efficiency of the engine 1 is even higher when the air-fuel ratio of the air-fuel mixture is lean. During SPCCI combustion, when the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio and when it is lean, the amount of fuel supplied to the cylinder is small and the amount of heat generated in the cylinder is small when the air-fuel ratio is lean. but less for lean. Since the amount of heat generated in the cylinder is small in the first place, when the air-fuel ratio of the air-fuel mixture is lean, the change in the temperature of the exhaust gas with respect to the retardation of the progress of combustion is small. Therefore, in the model 803 when the air-fuel ratio of the air-fuel mixture is lean, the temperature rise rate of the exhaust gas with respect to the change in the degree of combustion is lower than the temperature rise rate in the model 802 when the air-fuel ratio is stoichiometric.

The model is changed according to the rotation speed of the engine 1 . FIG. 9 illustrates models 901 and 903 when the number of revolutions of the engine 1 is high and models 902 and 904 when the number of revolutions is low. The upper diagram in FIG. 9 is a model for SI combustion, and the lower diagram is a model for SPCCI combustion.

As shown in the upper diagram of FIG. 9, the model 901 when the rotation speed of the engine 1 is high estimates the temperature of the exhaust gas higher than the model 902 when the rotation speed is low. It is configured. When the degree of combustion progress is the same, the estimation unit 106 estimates the temperature of the exhaust gas to be higher when the rotational speed of the engine 1 is high than when the rotational speed is low.

When the rotation speed of the engine 1 is high, the temperature of the exhaust gas becomes higher than when the rotation speed is low. This is because the higher the rotation speed of the engine 1, the higher the cylinder wall temperature due to the increase in the number of combustions per unit time, and the shorter the combustion time in the cylinder, which reduces the cooling loss.

By configuring the models 901 and 902 so that the temperature of the exhaust gas for the same degree of combustion progress is estimated to be higher when the rotation speed of the engine 1 is high than when the rotation speed is low, the estimation unit 106 can: Even if the rotation speed of the engine 1 changes, the temperature of the exhaust gas can be accurately estimated.

This is the same during SPCCI combustion as shown in the lower diagram of FIG. In other words, the model 903 when the rotation speed of the engine 1 is high estimates the exhaust gas temperature with respect to the same degree of combustion progress higher than the model 904 when the rotation speed is low. Even if the rotation speed of the engine 1 changes, the estimation unit 106 can accurately estimate the temperature of the exhaust gas.

Further, as shown in the upper diagram of FIG. 9, the model 901 when the engine speed is high has a higher exhaust gas temperature rise rate with respect to the change in the degree of combustion progress than the model 902 when the engine speed is low. is configured to be That is, model 901 has a steeper slope than model 902 .

As described above, when the rotation speed of the engine 1 is high, the cylinder wall temperature is high, and the difference between the in-cylinder temperature and the cylinder wall temperature is relatively small. When the rotation speed of the engine 1 is high, if the in-cylinder temperature changes as the degree of combustion progresses, the ratio of the amount of change in the in-cylinder temperature to the difference between the in-cylinder temperature during combustion and the cylinder wall temperature increases. . That is, if the degree of combustion progresses when the engine 1 rotates at a high speed, the cooling loss changes significantly, and the temperature of the exhaust gas also changes significantly. Conversely, when the rotational speed of the engine 1 is low, the cylinder wall temperature is low, and the difference between the in-cylinder temperature during combustion and the cylinder wall temperature is relatively large. When the rotation speed of the engine 1 is low, even if the degree of combustion progresses and the cylinder temperature changes, the ratio of the amount of change in the cylinder temperature to the difference between the cylinder temperature and the cylinder wall temperature is small. Changes in exhaust gas temperature are small.

By constructing the models 901 and 902 so that the temperature rise rate of the exhaust gas with respect to the change in the degree of combustion progresses when the rotation speed of the engine 1 is high than when the rotation speed is low, the estimating unit 106 can accurately estimate the temperature of the exhaust gas even if the rotation speed of the engine 1 changes.

Further, as shown in the lower diagram of FIG. 9, even during SPCCI combustion, the model 903 when the rotation speed of the engine 1 is high is the same as the model 903 when the temperature rise rate of the exhaust gas with respect to the change in the degree of combustion progresses. It is configured to be higher than 904. That is, model 903 has a steeper slope than model 904 .

Here, FIG. 10 illustrates the relationship between the rotation speed of the engine 1 and the estimated exhaust gas temperature at a certain engine load. The rotation speed N1 of the engine 1 is the rotation speed at which the SPCCI combustion and the SI combustion are switched in the operation map 401, as shown in FIG. Therefore, to the left of N1 in FIG. 10, engine 1 performs SPCCI combustion, and to the right of N1, engine 1 performs SI combustion.

When the rotation speed of the engine 1 is lower than N1, as the rotation speed increases, the degree of progress of SPCCI combustion is retarded, so the temperature of the exhaust gas gradually rises.

When the rotation speed of the engine 1 is N1, SPCCI combustion and SI combustion are switched. The degree of combustion progress in SI combustion advances compared to the degree of combustion progress in SPCCI combustion. Therefore, the temperature of the exhaust gas decreases when switching from SPCCI combustion to SI combustion, and increases when switching from SI combustion to SPCCI combustion.

Further, when the rotation speed of the engine 1 is equal to or higher than N1, as the rotation speed increases, the degree of progress of SI combustion is retarded, so the temperature of the exhaust gas gradually rises. However, the rate of increase in the exhaust gas temperature with respect to the increase in the rotational speed of the engine 1 is smaller during SI combustion than during SPCCI combustion. That is, the slope of the line in FIG. 10 is gentle. This is because during SI combustion, the amount of retardation of the degree of combustion progress is small with respect to the increase in the rotational speed of the engine 1 .

During SPCCI combustion, which greatly depends on the state of the temperature and pressure in the cylinder, it is susceptible to changes in the volume inside the cylinder during the expansion stroke. During SPCCI combustion, the retardation amount of the degree of combustion progress is large with respect to the increase in the engine speed. For this reason, during SPCCI combustion, the rate of increase in the temperature of the exhaust gas with respect to the increase in the engine speed is large. That is, the slope of the line in FIG. 10 is steep.

FIG. 11 illustrates models 1001 and 1003 when the load on the engine 1 is high and models 1002 and 1004 when the load is low. The model is changed according to the load of the engine 1. The upper diagram in FIG. 11 is a model for SI combustion, and the lower diagram is a model for SPCCI combustion.

As shown in the upper diagram of FIG. 11, the model 1001 when the load of the engine 1 is high is configured to estimate the exhaust gas temperature higher than the model 1002 when the load is low for the same degree of combustion progress. ing. When the degree of combustion progress is the same, the estimation unit 106 estimates the temperature of the exhaust gas to be higher when the load of the engine 1 is high than when the load is low.

When the load of the engine 1 is high, the amount of fuel supplied to the cylinder increases, and the amount of heat generated in the cylinder increases. By configuring the models 1001 and 1002 to estimate the temperature of the exhaust gas for the same degree of combustion, when the load of the engine 1 is high, it is higher than when the load is low. The temperature of the exhaust gas can be accurately estimated even if the load of the engine changes.

This is the same during SPCCI combustion as shown in the lower diagram of FIG. That is, the model 1003 when the load of the engine 1 is high is configured to estimate the exhaust gas temperature higher than the model 1004 when the load is low for the same degree of combustion progress. As a result, even if the load of the engine 1 changes during SPCCI combustion, the estimator 106 can accurately estimate the temperature of the exhaust gas.

Further, as shown in the upper diagram of FIG. 11, the model 1001 when the load of the engine 1 is high has a higher temperature rise rate of the exhaust gas with respect to the change in the degree of combustion progress than the model 1002 when the load is low. is configured to That is, model 1001 has a steeper slope than model 1002 .

When the load of the engine 1 is high, the fuel supply amount is large. When the load of the engine 1 is high, the amount of change in the indicated work is large when the degree of combustion progresses. As a result, the temperature of the exhaust gas also changes significantly. Conversely, when the load of the engine 1 is low, the amount of fuel supplied is small, so even if the degree of combustion progresses, the amount of change in indicated work is small. As a result, the change in temperature of the exhaust gas is small.

By configuring the models 1001 and 1002 so that the temperature rise rate of the exhaust gas with respect to the change in the degree of combustion progresses when the load of the engine 1 is high than when the load is low, the estimation unit 106 Even if the load of the engine 1 changes, the temperature of the exhaust gas can be accurately estimated.

Also, as shown in the lower diagram of FIG. 11, even during SPCCI combustion, the model 1003 when the load of the engine 1 is high has a rate of temperature rise of the exhaust gas with respect to the change in the degree of combustion progress, which is higher than that of the model 1004 when the load is low. is configured to be higher. That is, model 1003 has a steeper slope than model 1004 .

Returning to the flowchart of FIG. 7, in step S6, the estimation unit 106 estimates the temperature of the exhaust gas using a model (for example, model 801) corresponding to SI combustion. In step S9, the estimation unit 106 estimates the temperature of the exhaust gas using a model (for example, model 802) that corresponds to SPCCI combustion and the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio. Furthermore, in step S11, the estimating unit 106 estimates the temperature of the exhaust gas using a model (for example, the model 803) that corresponds to the SPCCI combustion and the air-fuel ratio of the air-fuel mixture is lean.

In step S<b>7 following step S<b>6 , the correction unit 107 of the ECU 10 corrects the exhaust gas temperature estimated in step S<b>6 according to the temperature of the engine 1 . Further, in step S10 following step S9, the correction unit corrects the temperature of the exhaust gas estimated in step S9 according to the temperature of the engine 1. FIG. Furthermore, in step S<b>12 following step S<b>11 , the correction unit corrects the exhaust gas temperature estimated in step S<b>11 according to the temperature of the engine 1 .

The temperature of the engine 1 is represented by the temperature of the engine 1 coolant. The correction unit 107 acquires the temperature of the cooling liquid of the engine 1 from the signal of the second liquid temperature sensor SW10.

FIG. 12 illustrates the relationship between the temperature of the coolant and the amount of correction for the temperature of the exhaust gas. T1 is the reference temperature of the coolant. All of the models 801, 802, 803, 901, 902, 903, 904, 1001, 1002, 1003, and 1004 described above describe the relationship between the degree of combustion progress and the exhaust gas temperature when the coolant is at the reference temperature T1. represents.

If the coolant temperature is T1, the correction amount is zero. The correction unit 107 does not actually correct the temperature of the exhaust gas estimated in steps S6, S9, and S11. If the coolant temperature is higher than T1, the correction amount is greater than zero. The higher the coolant temperature, the larger the positive correction amount. The correction unit 107 performs correction to raise the temperature of the exhaust gas estimated in steps S6, S9, and S11. Since the cooling loss decreases as the temperature of the engine 1 increases, the temperature of the exhaust gas increases. The ECU 10 can more accurately estimate the temperature of the exhaust gas by correcting the temperature of the exhaust gas by the correction unit 107 to raise the temperature of the exhaust gas.

If the coolant temperature is lower than T1, the correction amount is less than zero. The lower the coolant temperature, the larger the negative correction amount. The correction unit 107 performs correction to lower the exhaust gas temperature estimated in steps S6, S9, and S11. Since the cooling loss increases as the temperature of the engine 1 decreases, the temperature of the exhaust gas decreases. The ECU 10 can more accurately estimate the temperature of the exhaust gas by correcting the temperature of the exhaust gas by the correction unit 107 to lower the temperature of the exhaust gas.

The correction amount is changed according to the air-fuel ratio of the air-fuel mixture. Specifically, a straight line 1202 in FIG. 12 indicates the correction amount when SPCCI combustion is performed and the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio. A straight line 1203 indicates the correction amount when SPCCI combustion is performed and the air-fuel ratio of the air-fuel mixture is lean. As can be seen from FIG. 12, when the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio, the correction amount for the temperature of the exhaust gas is made larger than when it is lean.

When the air-fuel ratio of the air-fuel mixture is lean, the thermal efficiency of the engine 1 is relatively high, so the amount of fuel supplied into the cylinder is small. Therefore, when the air-fuel ratio of the air-fuel mixture is lean, the temperature inside the cylinder during combustion is lower than when the air-fuel ratio is stoichiometric. . Conversely, when the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio, the cooling loss increases.

The correction unit 107 performs correction in consideration of the influence of cooling loss. Therefore, when the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio, the correction amount for the temperature of the exhaust gas is made larger than when it is lean. Accordingly, the correction unit 107 can appropriately correct the temperature of the exhaust gas in consideration of the relationship between the air-fuel ratio of the air-fuel mixture and the cooling loss.

The correction amount is changed according to the combustion mode. Specifically, a straight line 1201 in FIG. 12 indicates the correction amount during SI combustion. As can be seen from FIG. 12, during SI combustion, the correction amount for the temperature of the exhaust gas is made larger than during SPCCI combustion.
In SPCCI combustion, the heat release rate during CI combustion is high, so the peak temperature in the cylinder is higher than the peak temperature during SI combustion. In SPCCI combustion, the difference between the peak temperature and the coolant temperature is large, so when the coolant temperature changes, the ratio of the coolant temperature change to the difference between the peak temperature and the coolant temperature is small. . Therefore, in SPCCI combustion, the cooling loss does not change much even if the coolant temperature changes. In SPCCI combustion, even if the temperature of the coolant changes, the temperature of the exhaust gas does not change significantly.

On the other hand, in SI combustion, the difference between the peak temperature and the coolant temperature is small, so when the coolant temperature changes, the ratio of the coolant temperature change to the difference between the peak temperature and the coolant temperature is big. In SI combustion, the cooling loss changes greatly when the coolant temperature changes. In SI combustion, the temperature of the exhaust gas changes greatly when the temperature of the coolant changes.

Therefore, during SI combustion, the correction amount for the temperature of the exhaust gas is made larger than during SPCCI combustion. Accordingly, the correction unit 107 can appropriately correct the exhaust gas temperature in consideration of the relationship between the combustion mode and the cooling loss.

Returning to the flowchart of FIG. 7, in step S7, the correction unit 107 corrects the temperature of the exhaust gas according to the correction amount 1201 corresponding to SI combustion. Further, in step S10, the correction unit 107 corrects the temperature of the exhaust gas according to the correction amount 1202 corresponding to the SPCCI combustion and the air-fuel ratio of the air-fuel mixture being the stoichiometric air-fuel ratio. Furthermore, in step S12, the correction unit 107 corrects the temperature of the exhaust gas according to the correction amount 1203 corresponding to the SPCCI combustion and the lean air-fuel ratio of the mixture.

The ECU 10, which has estimated the temperature of the exhaust gas according to the above procedure, performs control to lower the temperature of the exhaust gas when the estimated temperature of the exhaust gas is higher than the reference. The ECU 10 increases, for example, the amount of fuel supplied into the cylinder, compared to when the amount is below the standard. When the amount of fuel supplied into the cylinder is increased, the latent heat of the increased amount of fuel lowers the temperature of the exhaust gas discharged from the cylinder. By lowering the temperature of the exhaust gas below the standard, the reliability of the catalyst interposed in the exhaust passage 50 of the engine 1 can be ensured.

The ECU 10 may also lower the temperature of the coolant supplied to the engine 1 below the standard. Specifically, the ECU 10 controls the temperature of the coolant supplied to the engine 1 by controlling the thermostat valve 76 and the flow control valve 77 of the cooling system 70 . Thereby, the cooling loss of the engine 1 can be adjusted, and the temperature of the exhaust gas discharged from the cylinder can be lowered.

Further, the ECU 10 may increase the flow rate of the cooling liquid supplied to the engine 1 to a level below the standard. Specifically, the ECU 10 adjusts the flow rate of the coolant supplied to the engine 1 by controlling the thermostat valve 76 and the flow control valve 77 of the cooling system 70 . Thereby, the cooling loss of the engine 1 can be adjusted, and the temperature of the exhaust gas discharged from the cylinder can be lowered.

It should be noted that the control for lowering the temperature of the exhaust gas described above can be combined.

In the above configuration example, the ECU 10 estimates the temperature of the exhaust gas using a model representing the relationship between the degree of combustion progress and the temperature of the exhaust gas, and converts the estimated temperature of the exhaust gas into the engine 1 temperature. corrected accordingly. In contrast to this, a model representing the relationship between the degree of combustion progress, the exhaust gas temperature, and the engine temperature is created, and the ECU 10 uses this model to estimate the temperature of the exhaust gas from the degree of combustion progress and the engine temperature. You may This configuration corresponds to a configuration in which the estimation unit includes the correction unit.

In place of the model representing the relationship between the degree of combustion progress and the temperature of the exhaust gas, the ECU 10 estimates the temperature of the exhaust gas using at least a map representing the relationship between the degree of combustion progress and the temperature of the exhaust gas. good too.

Moreover, the technology disclosed herein is not limited to application to the engine 1 having the configuration described above. Various configurations can be adopted for the configuration of the engine 1 . For example, the technique disclosed herein may be applied to a diesel engine in which forced ignition is not performed on the air-fuel mixture in the cylinder. Even in a diesel engine, the control unit can accurately estimate the temperature of the exhaust gas from the model representing the relationship between the degree of combustion progress and the temperature of the exhaust gas, and the degree of combustion progress.

1 engine (engine)
10 ECU (control unit)
105 Calculation unit 106 Estimation unit 50 Exhaust passages 801, 802, 803 Models 901, 902, 903, 904 Models 1001, 1002, 1003, 1004 Model SW6 In-cylinder pressure sensor SW11 Crank angle sensor

Claims (11)

engine and
an exhaust passage that is connected to the engine and guides exhaust gas from a cylinder of the engine;
a sensor that outputs a signal corresponding to the combustion state in the cylinder;
a control unit connected to the sensor, estimating the temperature of the exhaust gas based on the signal of the sensor, and controlling the engine according to the estimated temperature of the exhaust gas,
The control unit performs forced ignition of the air-fuel mixture in the cylinder and burns the air-fuel mixture by flame propagation according to the operating state of the engine. and switching between a second combustion mode in which part of the air-fuel mixture is burned by self-ignition,
The control unit
a calculation unit that calculates a combustion progress degree, which is a crank angle at which a mass combustion ratio of combustion in the cylinder is 10%, 50%, or 90%, based on the signal from the sensor;
an estimation unit that estimates the temperature of the exhaust gas from the degree of combustion progress calculated by the calculation unit based on the relationship between the degree of combustion progress and the temperature of the exhaust gas;
The estimation unit estimates the temperature of the exhaust gas based on a first relationship between the degree of combustion progress and the temperature of the exhaust gas during the first combustion mode, and estimates the temperature of the exhaust gas during the second combustion mode. A controller for an engine that estimates the temperature of the exhaust gas based on a second relationship between the degree of combustion and the temperature of the exhaust gas, different from the first relationship. In the engine control device according to claim 1,
The estimating unit estimates the temperature of the exhaust gas to be higher in the first combustion mode than in the second combustion mode when the degree of progress of combustion is the same. In the engine control device according to claim 1 or 2,
In each of the first combustion mode and the second combustion mode, when the degree of combustion progress is retarded relative to the degree of combustion progress in the MBT, the estimating unit calculates a , an engine control device for linearly estimating the temperature of the exhaust gas. In the engine control device according to claim 3,
In each of the first combustion mode and the second combustion mode, when the degree of combustion progress is retarded from a knock limit, the estimating unit increases the retardation of the degree of combustion so that the exhaust gas An engine control device that linearly estimates the temperature of the engine. In the engine control device according to claim 3 or 4,
The control device for an engine, wherein the second relationship is such that the temperature rise rate of the exhaust gas with respect to the change in the degree of combustion progress is higher than the temperature rise rate of the first relationship. In the engine control device according to any one of claims 1 to 5,
In each of the first combustion mode and the second combustion mode, the estimating unit calculates the temperature of the exhaust gas for the same degree of combustion progress when the engine speed is high as compared to when the engine speed is low. High estimating engine control unit. In the engine control device according to claim 6,
In each of the first relationship and the second relationship, a temperature rise rate of the exhaust gas with respect to a change in the degree of combustion progress is higher when the engine speed is high than when the engine speed is low. In the engine control device according to any one of claims 1 to 7,
The estimating unit estimates the temperature of the exhaust gas for the same degree of combustion progress in each of the first combustion mode and the second combustion mode to be higher when the load of the engine is high than when the load is low. engine controller. In the engine control device according to claim 8,
In each of the first relationship and the second relationship, a temperature rise rate of the exhaust gas with respect to a change in the degree of combustion progress is higher when the load of the engine is high than when the load is low. In the engine control device according to any one of claims 1 to 9,
The engine control device, wherein the sensor is an in-cylinder pressure sensor that outputs a signal corresponding to the in-cylinder pressure. In the engine control device according to any one of claims 1 to 10,
The control unit for an engine, wherein when the temperature of the exhaust gas is higher than a reference, the amount of fuel supplied to the cylinder is increased more than when the temperature of the exhaust gas is equal to or lower than the reference.

Images