JP7031492B2 - How to design the control logic of a compression ignition engine - Google Patents

Description

The techniques disclosed herein relate to how to design the control logic of a compression ignition engine.

It is known that combustion by compression self-ignition, in which the air-fuel mixture burns at once without flame propagation, maximizes fuel efficiency because the combustion period is the shortest. However, combustion by compression self-ignition needs to solve various problems in an automobile engine. For example, in automobile applications, operating conditions and environmental conditions change significantly, so stable compression self-ignition is a major issue. Combustion by compression self-ignition has not yet been put into practical use in automobile engines. In order to solve this problem, for example, Patent Document 1 proposes that the spark plug ignites when the combustion chamber temperature is low and compression self-ignition is unlikely to occur. By igniting immediately before the compression top dead center, the pressure around the spark plug rises and compression self-ignition is promoted.

Japanese Patent No. 4082292

Unlike the technique described in Patent Document 1, which assists compression self-ignition by ignition of a spark plug, the applicant of the present application applies SPCCI (SPark Controlled Compression Ignition), which is a combination of SI (Spark Ignition) combustion and CI (Compression Ignition) combustion. ) Proposing combustion. SI combustion is combustion with flame propagation initiated by forcibly igniting the air-fuel mixture in the combustion chamber. CI combustion is combustion started by compression self-ignition of the air-fuel mixture in the combustion chamber. In SPCCI combustion, when the air-fuel mixture in the combustion chamber is forcibly ignited and combustion by flame propagation is started, the unburned air-fuel mixture in the combustion chamber is generated by the heat generation of SI combustion and the pressure increase due to flame propagation. Is a form of combustion by compression ignition. Since SPCCI combustion includes CI combustion, it is a form of "combustion by compression ignition".

CI combustion occurs when the in-cylinder temperature reaches the ignition temperature determined by the composition of the air-fuel mixture. If the in-cylinder temperature reaches the ignition temperature near the compression top dead center and CI combustion occurs, the fuel efficiency can be maximized. The in-cylinder temperature increases as the in-cylinder pressure increases. The in-cylinder pressure in SPCCI combustion is the result of two pressure increases: the pressure increase due to the compression work of the piston in the compression stroke and the pressure increase resulting from the heat generation of SI combustion.

Here, the compression work of the piston is determined by the effective compression ratio. If the effective compression ratio is too low, the pressure rise due to the compression work of the piston is small. In this case, the in-cylinder temperature cannot be raised to the ignition temperature unless the flame propagation in SPCCI combustion progresses and the pressure rise caused by the heat generated by SI combustion increases considerably. As a result, the amount of the air-fuel mixture that is compressed and self-ignited is small, and a large amount of the air-fuel mixture is burned by flame propagation, so that the combustion period is long and the fuel efficiency is lowered. That is, in order to stably cause CI combustion in SPCCI combustion and maximize fuel efficiency, it is necessary to keep the effective compression ratio at a certain value or higher.

On the other hand, if CI combustion occurs near the compression top dead center due to the high temperature inside the cylinder at the start of compression due to the high outside air temperature, the pressure inside the cylinder rises excessively and the combustion noise becomes excessive. become. In this case, for example, if the ignition timing is retarded, the combustion noise can be suppressed. However, if the ignition timing is retarded, CI combustion occurs when the piston is considerably lowered in the expansion stroke, which causes a decrease in fuel efficiency. In SPCCI combustion, the pressure increase generated from the heat generated by SI combustion can be used. Therefore, in order to achieve both suppression of combustion noise and improvement of fuel efficiency, the effective compression ratio is lowered to increase the pressure due to the compression work of the piston. It is effective to reduce. By doing so, it is possible to appropriately maintain the combustion noise without deteriorating the fuel efficiency.

In order to put an engine that performs SPCCI combustion into practical use, the designer determines the minimum effective compression ratio that stably produces CI combustion with respect to the operating condition of the engine, and the effective compression ratio as long as combustion noise is allowed. It is necessary to determine the effective compression ratio so that the combustion noise does not become excessive. By doing so, the designer can put into practical use an engine that maximizes fuel efficiency by maximizing the ratio of CI combustion in SPCCI combustion while keeping the combustion noise within an allowable range.

However, since SPCCI combustion is a new combustion method, no one has found an appropriate effective compression ratio range so far.

The effective compression ratio of the engine is determined by the geometric compression ratio and the intake valve closing timing. Since the intake valve closing timing is changed when the engine is actually operated, the designer needs to determine the range of the intake valve closing timing when designing the control logic of the engine that performs SPCCI combustion.

As the geometric compression ratio increases, the maximum intracranial pressure increases, so it is necessary to increase the strength of the engine components, which leads to an increase in weight and thus an increase in mechanical resistance loss. On the other hand, from the viewpoint of thermal efficiency, it is preferable that the expansion ratio determined from the geometric compression ratio is large. Since the maximum in-cylinder pressure changes depending on the heat account based on the combustion form and stroke volume even if the geometric compression ratio is the same, the optimum geometric compression ratio for SPCCI combustion, which is a new combustion method, has been so far. The place is unknown.

As a result of diligent studies on SPCCI combustion, the present inventors have succeeded in finding an appropriate intake valve closing timing range within the range of geometric compression ratios at which SPCCI combustion can occur. Based on this finding, the inventors of the present invention have invented a method for designing a control logic for a compression ignition type engine.

Specifically, the technique disclosed herein relates to a method of designing a control logic of a compression ignition type engine mounted on an automobile .

The engine has a fuel injection unit that injects fuel to be supplied into a combustion chamber partitioned by a cylinder and a piston, a variable valve mechanism that changes the valve timing of an intake valve that opens and closes the combustion chamber, and a combustion chamber. The ignition unit that ignites the air-fuel mixture, the measurement unit that measures parameters related to the operating state of the engine, and the measurement results of the measuring unit are received, and calculations are performed according to the control logic corresponding to the operating state of the engine. , The fuel injection unit, the variable valve mechanism, and a control unit that outputs a signal to the ignition unit.

In the control unit, between a predetermined light load and an upper limit load, the G / F, which is the weight ratio of the total gas of the air-fuel mixture in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. A signal is output to each of the fuel injection unit and the variable valve mechanism so that the A / F, which is the weight ratio of the air-fuel mixture to the fuel, becomes richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio. After the ignition unit ignites the air-fuel mixture in the combustion chamber, a signal is output to the ignition unit so that the unburned air-fuel mixture burns by self-ignition.

The method of designing the control logic includes a step of determining the geometric compression ratio ε of the engine and a step of determining the control logic that determines the valve closing timing IVC of the intake valve, and sets the control logic. If the fuel is a high octane fuel and the geometric compression ratio ε is 10 ≦ ε <17.
0.4234ε ^2-22.926ε ＋ 207.84 ＋ C ≦ IVC ≦ －0.4234ε ² ＋ 22.926ε－167.84 ＋ C… (a)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.
However, C is a correction term related to the engine speed NE (rpm), and C = 3.3 × 10 ^-10 NE ³ − 1.0 × 10 ^-6 NE ² + 7.0 × 10 ^-4 NE.

The ignition unit receives a signal from the control unit and ignites the air-fuel mixture in the combustion chamber. Combustion by flame propagation starts, and then the unburned air-fuel mixture burns by self-ignition to complete the combustion. That is, this engine performs SPCCI combustion.

The engine makes the G / F of the air-fuel mixture leaner than the stoichiometric air-fuel ratio and makes the A / F richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio. By making the G / F of the air-fuel mixture lean, the fuel efficiency of the engine is improved. By setting the A / F to the stoichiometric air-fuel ratio, the exhaust emission performance can be improved by utilizing the catalyst device.

When designing the control logic of an engine, the designer first determines the geometric compression ratio ε of the engine. When the geometric compression ratio ε is set to 10 ≦ ε <17, the designer determines the valve closing timing IVC of the intake valve so as to satisfy the above equation (a). By determining the valve closing timing so as to satisfy the equation (a), the engine can perform SI combustion and CI combustion while keeping the combustion noise within an allowable range even under various conditions in which the conditions in the combustion chamber are different. It becomes possible to perform stable CI combustion in the combined SPCCI combustion. By operating the engine that performs SPCCI combustion according to this control logic, the fuel efficiency is maximized.

So far, there has been no index for determining the valve closing timing IVC of the intake valve that can be used when designing the control logic of the engine that performs SPCCI combustion. The designer had to determine the IVC corresponding to the operating state of the engine by repeating experiments and the like under various conditions while changing the valve closing timing IVC of the intake valve to various timings.

The above design method defines the relationship between the geometric compression ratio ε of the engine and the valve closing timing IVC of the intake valve in order to realize appropriate SPCCI combustion. The designer can put into practical use an engine that performs SPCCI combustion if the valve closing timing IVC of the intake valve is determined within the range that satisfies the relationship. The designer can put an engine that performs SPCCI combustion into practical use with extremely few man-hours.

When the control logic is set, if the fuel is a high octane fuel and the geometric compression ratio ε is 17 ≦ ε <20.
－0.4288ε ² ＋31.518ε－379.88 ＋ C ≦ IVC ≦ －0.4234ε ² ＋22.926ε－167.84 ＋ C… (b)
Or,
0.4234ε 2-22.926ε ＋ 207.84 ＋ C ≦ IVC ≦ 1.9163ε ^{2 －89.935ε} ＋ 974.94 ＋ C… (c)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

As a result, the fuel efficiency of the engine that performs SPCCI combustion is maximized. Further, the designer can put into practical use an engine that performs SPCCI combustion with a small number of man-hours.

When the control logic is set, if the fuel is a high octane fuel and the geometric compression ratio ε is 20 ≦ ε ≦ 30.
－0.4288ε ² ＋31.518ε－379.88 ＋ C ≦ IVC ≦ 120… (d)
Or,
-80 ≤ IVC ≤ 1.9163ε ² -89.935ε + 974.94 + C ... (e)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Similar to the above, the engine that performs SPCCI combustion has the maximum fuel efficiency. Further, the designer can put into practical use an engine that performs SPCCI combustion with a small number of man-hours.

When the control logic is set, if the fuel is a low octane fuel and the geometric compression ratio ε is 10 ≦ ε <15.7.
0.4234ε ² －21.826ε ＋ 178.75 ＋ C ≦ IVC ≦ －0.4234ε ² ＋ 21.826ε－138.75 ＋ C… (f)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

As a result, the fuel efficiency of an engine using a low octane number fuel is maximized. Further, the designer can put into practical use an engine that performs SPCCI combustion with a small number of man-hours.

When the geometric compression ratio ε is 10 ≦ ε <15.7, the designer determines the valve closing timing IVC of the intake valve so as to satisfy both the equation (a) and the equation (f). It is possible to set a control logic suitable for both an engine using a high octane fuel and an engine using a low octane fuel. Even if the octane number of the fuel differs depending on the destination, the designer can collectively design the engine control logic, and the design man-hours are reduced.

When the control logic is set, if the fuel is a low octane fuel and the geometric compression ratio ε is 15.7 ≦ ε <18.7.
-0.5603ε ² + 34.859ε-377.22 + C ≤ IVC ≤ -0.4234ε ² + 21.826ε-138.75 + C ... (g)
Or,
0.4234ε ² －21.826ε ＋ 178.75 ＋ C ≦ IVC ≦ 1.9211ε ² －85.076ε ＋ 862.01 ＋ C… (h)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

When the control logic is set, if the fuel is a low octane fuel and the geometric compression ratio ε is 18.7 ≦ ε ≦ 30.
-0.5603ε ² + 34.859ε-377.22 + C ≤ IVC ≤ 120 ... (i)
Or,
-80 ≤ IVC ≤ 1.9211ε ^2-85.076ε + 862.01 + C ... (j)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

The valve closing timing IVC of the intake valve changes when the operating state of the engine changes, and the closing timing is such that the relational expression of any one of (a) to (j) is satisfied for each operating state. The valve timing IVC (deg. ABDC) may be determined.

By doing so, the engine can stably perform SPCCI combustion while suppressing combustion noise in various operating conditions.

The upper limit load may be a high load of a predetermined load or more, and the engine may be operated in a high load operation state of a predetermined load or more according to the control logic.

In general CI combustion, the pressure fluctuation at the time of ignition is relatively large, so when the load on the engine becomes high, the combustion noise becomes too large and exceeds the permissible value. On the other hand, SPCCI combustion is SI combustion at the start of combustion. In SI combustion, the pressure fluctuation at the time of ignition is small. SPCCI combustion can suppress combustion noise even when the load on the engine is high.

The upper limit load may be the maximum load, and the engine may be operated in the maximum load operating state according to the control logic. That is, the engine may perform SPCCI combustion in the maximum load operating state.

The engine includes a supercharger, and the control unit causes the supercharger to supercharge when the engine is in a high load operation state of a predetermined load or more according to the control logic. It may output a signal.

When the load on the engine increases and the amount of fuel supplied to the combustion chamber increases, the turbocharger supercharges the mixture to make the G / F of the air-fuel mixture leaner than the stoichiometric air-fuel ratio and the A / F. Is realized to be richer than the theoretical air-fuel ratio or the theoretical air-fuel ratio. It is advantageous for improving the fuel efficiency of the engine.

The control unit may set the G / F of the air-fuel mixture to 18 or more and 50 or less according to the control logic.

The engine includes an EGR system that introduces exhaust gas into the combustion chamber, and the control unit uses flame propagation to propagate the total amount of heat generated when the air-fuel mixture in the combustion chamber burns. A signal is output to the EGR system and the ignition unit so that the calorific value ratio as an index related to the ratio of the calorific value generated when the engine burns becomes the target calorific value ratio determined according to the operating state of the engine. May be done.

SPCCI combustion has a calorific value ratio of less than 100%. In the combustion mode (that is, SI combustion) in which combustion is completed only by combustion by flame propagation without generating combustion by compression ignition, the calorific value ratio is 100%.

When the calorific value ratio is increased in SPCCI combustion, the ratio of SI combustion is increased, which is advantageous in suppressing combustion noise. When the calorific value ratio is lowered in SPCCI combustion, the ratio of CI combustion is increased, which is advantageous for improving fuel efficiency. The calorific value ratio changes by changing the temperature in the combustion chamber and / or the ignition timing. For example, when the temperature in the combustion chamber is high, the start timing of CI combustion is early, so that the calorific value ratio is low. Further, when the ignition timing is advanced, the SI combustion start timing is earlier, so that the calorific value ratio becomes higher. By outputting a signal to the EGR system and the ignition unit so that the calorific value ratio becomes the target calorific value ratio determined according to the operating state of the engine, it is possible to suppress combustion noise and improve fuel efficiency at the same time. ..

The control unit may output a signal to the EGR system or the ignition unit when the load of the engine is high so that the calorific value ratio is higher than when the load is low.

When the load on the engine increases, the amount of fuel supplied to the combustion chamber increases and the temperature inside the combustion chamber rises. Combustion noise can be suppressed by increasing the calorific value ratio of SPCCI combustion when the load of the engine is high.

As described above, according to the method of designing the control logic of the compression ignition type engine, it is possible to design the control logic that realizes SPCCI combustion that maximizes the fuel efficiency.

FIG. 1 is a diagram illustrating an engine configuration. FIG. 2 is a diagram illustrating the configuration of the combustion chamber, the upper figure is a plan view equivalent view of the combustion chamber, and the lower figure is a sectional view taken along line II-II. FIG. 3 is a plan view illustrating the configuration of the combustion chamber and the intake system. FIG. 4 is a block diagram illustrating the configuration of the engine control device. FIG. 5 is a diagram illustrating a waveform of SPCCI combustion. FIG. 6 is a diagram illustrating an engine map, the upper figure is a warm map, the middle map is a semi-warm map, and the lower figure is a cold map. FIG. 7 is a diagram illustrating the details of the map during warm weather. FIG. 8 is a diagram illustrating fuel injection timing, ignition timing, and combustion waveform in each operating region of the map of FIG. 7. FIG. 9 is a diagram illustrating the layer structure of the engine map. FIG. 10 is a flowchart illustrating a control process related to layer selection of a map. The upper figure of FIG. 11 is a diagram illustrating the relationship between the engine load and the valve opening timing of the intake valve in layer 2, and the lower figure is a diagram illustrating the relationship between the engine speed and the valve opening timing of the intake valve in layer 2. The upper figure of FIG. 12 shows the relationship between the engine load in layer 3 and the valve opening timing of the intake valve, the middle figure shows the relationship between the engine load in layer 3 and the valve closing timing of the exhaust valve, and the lower figure shows the engine in layer 3. It is a figure which illustrates the relationship between a load and an overlap period of an intake valve and an exhaust valve. FIG. 13 is a flowchart illustrating the process of engine operation control executed by the ECU. FIG. 14 is a diagram illustrating the relationship between the high and low of the engine load and the high and low of the target SI rate. FIG. 15 is a diagram showing a range in which SPCCI combustion is established with respect to the EGR rate in layer 2. FIG. 16 is an example of a matrix image used in layer 2 to determine the relationship between the geometric compression ratio capable of SPCCI combustion and the valve closing timing of the intake valve. The upper figure of FIG. 17 shows the relationship between the geometric compression ratio capable of SPCCI combustion in layer 2 and the valve closing timing of the intake valve when a high octane number fuel is used, and the lower figure shows the relationship between the valve closing timing of the intake valve and the lower figure when a low octane number fuel is used. This is the relationship between the geometric compression ratio at which SPCCI combustion is possible in layer 2 and the valve closing timing of the intake valve. FIG. 18 is a diagram showing a range in which SPCCI combustion is stable with respect to G / F in layer 3. FIG. 19 is an example of a matrix image used in layer 3 to determine the relationship between the geometric compression ratio capable of SPCCI combustion and the valve closing timing of the intake valve. The upper figure of FIG. 20 shows the relationship between the geometric compression ratio that enables SSPCCI combustion in layer 3 and the valve closing timing of the intake valve when a high octane number fuel is used, and the lower figure shows the relationship between the valve closing timing of the intake valve and the lower figure when a low octane number fuel is used. This is the relationship between the geometric compression ratio at which SPCCI combustion is possible in layer 3 and the valve closing timing of the intake valve. The upper figure of FIG. 21 shows the relationship between the geometric compression ratio capable of SPCCI combustion in layer 2 and layer 3 and the valve closing timing of the intake valve when a high octane number fuel is used, and the lower figure shows a low octane number fuel. It is the relationship between the geometric compression ratio at which SPCCI combustion is possible in layer 2 and layer 3 and the valve closing timing of the intake valve when the above is used. FIG. 22 is a flowchart illustrating the procedure of the method of designing the control logic of the compression ignition type engine.

Hereinafter, embodiments relating to a method for designing the control logic of the compression ignition type engine will be described in detail with reference to the drawings. The following description is an example of the design method of the engine and the control logic.

FIG. 1 is a diagram illustrating the configuration of a compression ignition type engine. FIG. 2 is a diagram illustrating the configuration of the combustion chamber of the engine. FIG. 3 is a diagram illustrating the configuration of the combustion chamber and the intake system. The intake side in FIG. 1 is on the left side of the paper surface, and the exhaust side is on the right side of the paper surface. The intake side in FIGS. 2 and 3 is on the right side of the paper, and the exhaust side is on the left side of the paper. FIG. 4 is a block diagram illustrating the configuration of the engine control device.

The engine 1 is a 4-stroke engine in which the combustion chamber 17 is operated by repeating an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke. The engine 1 is mounted on a four-wheeled vehicle. By driving the engine 1, the automobile runs. The fuel of the engine 1 is gasoline in this configuration example. The fuel may be at least a liquid fuel containing gasoline. The fuel may be gasoline containing, for example, bioethanol or the like.

(Engine configuration)
The engine 1 includes a cylinder block 12 and a cylinder head 13 mounted on the cylinder block 12. A plurality of cylinders 11 are formed inside the cylinder block 12. 1 and 2 show only one cylinder 11. The engine 1 is a multi-cylinder engine.

A piston 3 is slidably inserted in each cylinder 11. The piston 3 is connected to the crankshaft 15 via a connecting rod 14. The piston 3 partitions the combustion chamber 17 together with the cylinder 11 and the cylinder head 13. The "combustion chamber" may be used in a broad sense. That is, 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.

As shown in the lower figure of FIG. 2, the lower surface of the cylinder head 13, that is, the ceiling surface of the combustion chamber 17, is composed of an inclined surface 1311 and an inclined surface 1312. The inclined surface 1311 has an upward slope from the intake side toward the injection axis X2 of the injector 6, which will be described later. The inclined surface 1312 has an upward slope from the exhaust side toward the injection axis X2. The ceiling surface of the combustion chamber 17 has a so-called pent roof shape.

The upper surface of the piston 3 is raised toward the ceiling surface of the combustion chamber 17. A cavity 31 is formed on the upper surface of the piston 3. The cavity 31 is recessed from the upper surface of the piston 3. The cavity 31 has a shallow dish shape in this configuration example. The center of the cavity 31 is displaced toward the exhaust side from the central axis X1 of the cylinder 11.

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 a part of the operating region. SPCCI combustion controls CI combustion by utilizing heat generation and pressure increase due to SI combustion. The engine 1 is a compression ignition type engine. However, this engine 1 does not need to raise 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 set the geometric compression ratio relatively low. Lowering the geometric compression ratio is advantageous in reducing the cooling loss and the mechanical loss. The geometric compression ratio of engine 1 is 14 to 17 for regular specifications (low octane number fuel with an octane number of about 91), and 15 for high-octane specifications (high octane number fuel with an octane number of about 96). It may be ~ 18.

The cylinder head 13 is formed with an intake port 18 for each cylinder 11. As shown in FIG. 3, the intake port 18 has a first intake port 181 and a second intake port 182. The intake port 18 communicates with the combustion chamber 17. The intake port 18 is a so-called tumble port, although detailed illustration is omitted. That is, the intake port 18 has a shape such that a tumble flow is formed in the combustion chamber 17.

An intake valve 21 is provided at 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 operating mechanism. The valve operation mechanism may be a variable valve mechanism that makes the valve timing and / or the valve lift variable. In this configuration example, as shown in FIG. 4, the variable valve mechanism has an intake electric S-VT (Sequential-Valve Timing) 23. The intake electric S-VT23 continuously changes the rotation phase of the intake camshaft within a predetermined angle range. The valve opening timing and valve closing timing of the intake valve 21 change continuously. 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 191 and a second exhaust port 192, as shown in FIG. The exhaust port 19 communicates with the combustion chamber 17.

An exhaust valve 22 is provided at 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 operating mechanism. This valve operation mechanism may be a variable valve mechanism that makes the valve timing and / or the valve lift variable. In this configuration example, as shown in FIG. 4, the variable valve mechanism has an exhaust electric S-VT24. The exhaust electric S-VT24 continuously changes the rotational phase of the exhaust camshaft within a predetermined angular range. The valve opening timing and valve closing timing of the exhaust valve 22 change continuously. The exhaust valve mechanism may have a hydraulic S-VT instead of the electric S-VT.

The intake electric S-VT23 and the exhaust electric S-VT24 adjust the length of the overlap period in which both the intake valve 21 and the exhaust valve 22 are opened. By increasing the length of the overlap period, the residual gas in the combustion chamber 17 can be scavenged. Further, by adjusting the length of the overlap period, the internal EGR (Exhaust Gas Recirculation) gas can be introduced into the combustion chamber 17. The internal EGR system is composed of an intake electric S-VT23 and an exhaust electric S-VT24. The internal EGR system is not always composed of S-VT.

An injector 6 is attached to the cylinder head 13 for each cylinder 11. The injector 6 injects fuel directly into the combustion chamber 17. The injector 6 is an example of a fuel injection unit. The injector 6 is arranged in the valley portion of the pent roof where the inclined surface 1311 and the inclined surface 1312 intersect. As shown in FIG. 2, the injection shaft center X2 of the injector 6 is located on the exhaust side of the central shaft X1 of the cylinder 11. The injection axis X2 of the injector 6 is parallel to the central axis X1. The injection axis X2 of the injector 6 and the center of the cavity 31 coincide with each other. The injector 6 faces the cavity 31. The injection axis X2 of the injector 6 may coincide with the central axis X1 of the cylinder 11. In the case of that configuration, the injection axis X2 of the injector 6 and the center of the cavity 31 may coincide with each other.

Although detailed illustration is omitted, the injector 6 is composed of a multi-injection type fuel injection valve having a plurality of injection ports. The injector 6 injects fuel so that the fuel spray radiates from the center of the combustion chamber 17, as shown by the alternate long and short dash line in FIG. In the present configuration example, the injector 6 has ten injection holes, and the injection holes are arranged at equal angles in the circumferential direction.

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 that connects 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. The fuel pump 65 pumps fuel to the common rail 64. In this configuration example, the fuel pump 65 is a plunger type pump driven by a crankshaft 15. The common rail 64 stores the fuel pumped from the fuel pump 65 at a 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 fuel having a high pressure of 30 MPa or more to the injector 6. The pressure of the fuel supplied to the injector 6 may be changed according to the operating state of the engine 1. The configuration of the fuel supply system 61 is not limited to the above configuration.

A spark plug 25 is attached to the cylinder head 13 for each cylinder 11. The spark plug 25 forcibly ignites the air-fuel mixture in the combustion chamber 17. In this configuration example, the spark plug 25 is arranged on the intake side of the central axis X1 of the cylinder 11. The spark plug 25 is located between the two intake ports 18. The spark plug 25 is attached to the cylinder head 13 so as to be inclined from above to below toward the center of the combustion chamber 17. As shown in FIG. 2, the electrode of the spark plug 25 faces the inside of the combustion chamber 17 and is located near the ceiling surface of the combustion chamber 17. The spark plug 25 may be arranged on the exhaust side of the central axis X1 of the cylinder 11. Further, the spark plug 25 may be arranged on the central axis X1 of the cylinder 11.

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. The gas introduced into the combustion chamber 17 flows through the intake passage 40. An air cleaner 41 is disposed 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. The 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 degree of the valve.

The intake passage 40 is also provided with a supercharger 44 downstream of the throttle valve 43. The supercharger 44 supercharges the gas to be introduced into the combustion chamber 17. In this configuration example, the turbocharger 44 is a mechanical turbocharger driven by the engine 1. The mechanical turbocharger 44 may be a roots type, a Rishorum type, a vane type, or a centrifugal type.

An electromagnetic clutch 45 is interposed between the supercharger 44 and the engine 1. The electromagnetic clutch 45 transmits a driving force from the engine 1 to the supercharger 44 or cuts off the transmission of the driving force between the supercharger 44 and the engine 1. As will be described later, the turbocharger 44 is switched on and off by switching the shutoff and connection of the electromagnetic clutch 45 by the ECU 10.

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

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 to each other so as to bypass the supercharger 44 and the intercooler 46. An air bypass valve 48 is provided in the bypass passage 47. The air bypass valve 48 regulates 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 disengaged). The 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 is operated in a non-supercharged state, that is, in a naturally aspirated state.

When the supercharger 44 is turned on, the engine 1 operates in the 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). A part of the gas that has passed through the turbocharger 44 flows back to the upstream of the turbocharger 44 through the bypass passage 47. When the ECU 10 adjusts the opening degree of the air bypass valve 48, the boost pressure of the gas introduced into the combustion chamber 17 changes. The supercharging time is defined as the time when the pressure in the surge tank 42 exceeds the atmospheric pressure, and the non-supercharging time is defined as the time when the pressure in the surge tank 42 becomes the atmospheric pressure or less. May be good.

In this configuration example, the supercharging system 49 is configured by the supercharger 44, the bypass passage 47, and the air bypass valve 48.

The engine 1 has a swirl generation unit that generates a swirl flow in the combustion chamber 17. As shown in FIG. 3, the swirl generator has a swirl control valve 56 attached to the intake passage 40. The swirl control valve 56 is arranged in the secondary passage 402 of the primary passage 401 connected to the first intake port 181 and the secondary passage 402 connected to the second intake port 182. The swirl control valve 56 is an opening degree controlling valve capable of narrowing the cross section of the secondary passage 402. When the opening degree of the swirl control valve 56 is small, the intake flow flow from the first intake port 181 to the combustion chamber 17 is relatively large, and the intake flow flow from the second intake port 182 to the combustion chamber 17 is relatively large. Since the amount is small, the swirl flow in the combustion chamber 17 becomes strong. When the opening degree of the swirl control valve 56 is large, the intake flow rates flowing into the combustion chamber 17 from each of the first intake port 181 and the second intake port 182 become substantially equal, so that the swirl flow in the combustion chamber 17 is weak. Become. When the swirl control valve 56 is fully opened, no swirl flow is generated. The swirl flow orbits in the counterclockwise direction in FIG. 3 as shown by the white arrow (see also the white arrow in FIG. 2).

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. Although not shown in detail, the upstream portion of the exhaust passage 50 constitutes an independent passage that branches for each cylinder 11. The upstream end of the independent passage is connected to the exhaust port 19 of each cylinder 11.

An exhaust gas purification system having 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. The downstream catalytic converter is located outside the engine room. The downstream catalytic converter has a three-way catalyst 513. The exhaust gas purification system is not limited to the configuration shown in the figure. For example, GPF may be omitted. Further, the catalytic converter is not limited to the one having a three-way catalyst. Further, the order of the three-way catalyst and the GPF may be changed as appropriate.

An EGR passage 52 constituting an external EGR system is connected between the intake passage 40 and the exhaust passage 50. The EGR passage 52 is a passage for returning a part of the exhaust gas to the intake passage 40. The upstream end of the EGR passage 52 is connected between the upstream catalytic converter and the downstream catalytic converter in the exhaust passage 50. The downstream end of the EGR passage 52 is connected to the upstream portion of the turbocharger 44 in the intake passage 40. The EGR gas flowing through the EGR passage 52 enters the upstream portion of the turbocharger 44 in the intake passage 40 without passing through the air bypass valve 48 of the bypass passage 47.

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

In 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 the combustion chamber 17 with exhaust gas having a lower temperature than the internal EGR system.

The control device of the compression ignition type engine includes an ECU (Engine Control Unit) 10 for operating the engine 1. The ECU 10 is a controller based on a well-known microcomputer, and as shown in FIG. 4, a central processing unit (CPU) 101 for executing a program, and for example, a RAM (Random Access Memory) or a ROM. It includes a memory 102 configured by (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 unit.

As shown in FIGS. 1 and 4, various sensors SW1 to SW17 are connected to the ECU 10. The sensors SW1 to SW17 output a signal to the ECU 10. The sensors include the following sensors.

Air flow sensor SW1: Arranged downstream of the air cleaner 41 in the intake passage 40 and measuring the flow rate of fresh air flowing through the intake passage 40. First intake temperature sensor SW2: Arranged downstream of the air cleaner 41 in the intake passage 40. First pressure sensor SW3 that measures the temperature of fresh air flowing through the intake passage 40: It is located downstream of the connection position of the EGR passage 52 in the intake passage 40 and upstream of the supercharger 44, and is the supercharger 44. Second intake temperature sensor SW4 that measures the pressure of the gas flowing into the engine: It is located downstream of the turbocharger 44 in the intake passage 40 and upstream of the connection position of the bypass passage 47, and flows out of the turbocharger 44. Second pressure sensor SW5 that measures the temperature of the gas: Attached to the surge tank 42 and finger pressure sensor SW6 that measures the pressure of the gas downstream of the turbocharger 44: Attached to the cylinder head 13 corresponding to each cylinder 11. Exhaust temperature sensor SW7 that measures the pressure in each combustion chamber 17: Linear _O2 sensor SW8 that is arranged in the exhaust passage 50 and measures the temperature of the exhaust gas discharged from the combustion chamber 17: upstream in the exhaust passage 50. Lambda O ₂ sensor SW9 located upstream of the catalytic converter and measuring the oxygen concentration in the exhaust gas: Arranged downstream of the ternary catalyst 511 in the upstream catalytic converter and measuring the oxygen concentration in the exhaust gas. Water temperature sensor SW10: Attached to the engine 1 and measures the temperature of the cooling water Crank angle sensor SW11: Attached to the engine 1 and measures the rotation angle of the crankshaft 15 Accelerator opening sensor SW12: Attached to the accelerator pedal mechanism Intake cam angle sensor SW13: attached to the engine 1 and exhaust cam angle sensor SW14: attached to the engine 1 to measure the rotation angle of the intake cam shaft. EGR differential pressure sensor SW15 that measures the rotation angle of the exhaust cam shaft: is arranged in the EGR passage 52, and fuel pressure sensor SW16 that measures the differential pressure upstream and downstream of the EGR valve 54: on the common rail 64 of the fuel supply system 61. Third intake temperature sensor SW17 that is attached and measures the pressure of the fuel supplied to the injector 6: The temperature of the gas that is attached to the surge tank 42 and is introduced into the combustion chamber 17, in other words, the intake air that is introduced into the combustion chamber 17. Measure the temperature.

The ECU 10 determines the operating state of the engine 1 based on the signals of the sensors SW1 to SW17, and calculates the control amount of each device according to a predetermined control logic. The control logic is stored in the memory 102. The control logic includes calculating a target amount and / or a control amount using a map stored in the memory 102.

The ECU 100 outputs an electric signal related to the calculated control amount to the injector 6, the spark plug 25, the intake electric S-VT23, the exhaust electric S-VT24, the fuel supply system 61, the throttle valve 43, the EGR valve 54, and the supercharger 44. It outputs to the electromagnetic clutch 45, the air bypass valve 48, and the swirl control valve 56.

For example, the ECU 10 sets the target torque of the engine 1 and determines the target boost pressure based on the signal of the accelerator opening sensor SW12 and the map. Then, the ECU 10 adjusts the opening degree of the air bypass valve 48 based on the target boost pressure and the front-rear differential pressure of 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 becomes the target supercharging pressure.

Further, the ECU 10 sets a target EGR ratio (that is, a ratio of EGR gas to all gas in the combustion chamber 17) based on the operating state of the engine 1 and the map. Then, 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 the front-rear differential pressure of the EGR valve 54 obtained from the signal of the EGR differential pressure sensor SW15. By performing feedback control for adjusting the opening degree of the EGR valve 54 based on the above, the amount of external EGR gas introduced into the combustion chamber 17 is set to be the target EGR gas amount.

Further, the ECU 10 executes the air-fuel ratio feedback control when a predetermined control condition is satisfied. Specifically, the ECU 10 injects fuel from 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 O ₂ sensor SW8 and the lambda _O2 sensor SW9. Adjust the amount.

The details of the control of the engine 1 by the other ECU 10 will be described later.

(SPCCI combustion concept)
The engine 1 performs combustion by compression self-ignition under a predetermined operating state for the main purpose of improving fuel efficiency and exhaust gas performance. In combustion by self-ignition, when the temperature in the combustion chamber 17 before the start of compression fluctuates, the timing of self-ignition changes significantly. Therefore, the engine 1 performs SPCCI combustion that combines SI combustion and CI combustion.

In SPCCI combustion, the ignition plug 25 forcibly ignites the air-fuel mixture in the combustion chamber 17, so that the air-fuel mixture undergoes SI combustion by flame propagation, and the heat generated by the SI combustion causes SI combustion in the combustion chamber 17. This is a form in which the unburned air-fuel mixture undergoes CI combustion by self-ignition as the temperature rises and the pressure in the combustion chamber 17 rises due to flame propagation.

By adjusting the calorific value of SI combustion, it is possible to absorb the variation in temperature 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 generation during SI combustion is milder than the heat generation during CI combustion. As illustrated in FIG. 5, the waveform of the heat generation rate in the SPCCI combustion has a rising slope smaller than the rising 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 waveform of the heat generation rate may change from small to large at the timing of self-ignition. The waveform of the heat generation rate may have an inflection point X at the timing when CI combustion starts.

After the start of CI combustion, SI combustion and CI combustion are performed in parallel. Since CI combustion generates more heat than SI combustion, the heat generation rate is relatively large. However, since CI combustion is performed after the compression top dead center, it is avoided that the slope of the heat generation rate waveform becomes too large. The pressure fluctuation (dp / dθ) during CI combustion is also relatively gentle.

The pressure fluctuation (dp / dθ) can be used as an index representing the combustion noise. As described above, in the SPCCI combustion, the pressure fluctuation (dp / dθ) can be reduced, so that it is possible to avoid the combustion noise becoming too large. The combustion noise of the engine 1 is suppressed to an allowable level or less.

When the CI combustion ends, the SPCCI combustion ends. CI combustion has a shorter combustion period than SI combustion. SPCCI combustion has an earlier combustion end time than SI combustion.

The heat generation rate waveform of SPCCI combustion is formed so that the first heat generation rate part Q _SI formed by SI combustion and the second heat generation part Q _CI formed by CI combustion are continuous in this order. Has been done.

Here, the SI rate is defined as a parameter indicating the characteristics of SPCCI combustion. The SI rate is defined as an index related to the ratio of the amount of heat generated by SI combustion to the total amount of heat generated by SPCCI combustion. The SI rate is the ratio of the amount of heat generated by two types of combustion having different combustion forms. When the SI rate is high, the rate of SI combustion is high, and when the SI rate is low, the rate of CI combustion is high. The SI rate may be defined as the ratio of the amount of heat generated by SI combustion to the amount of heat generated by CI combustion. That is, in the waveform 801 shown in FIG. 5, SI rate = (SI combustion area: Q _SI ) / (CI combustion area: Q _CI ).

When the engine 1 performs SPCCI combustion, a strong swirl flow is generated in the combustion chamber 17. A strong swirl flow may be defined as, for example, a flow having a swirl ratio of 4 or more. The swirl ratio can be defined as a value obtained by measuring and integrating the intake flow lateral angular velocity for each valve lift and dividing it by the engine angular velocity. Although not shown, the intake flow lateral angular velocity can be determined based on the measurement using a known rig test device.

When a strong swirl flow is generated in the combustion chamber 17, the outer peripheral portion of the combustion chamber 17 becomes a strong swirl flow, while the swirl flow in the central portion becomes relatively weak. Due to the vortex flow caused by the velocity gradient at the boundary between the central portion and the outer peripheral portion, the turbulent energy becomes high in the central portion. When the spark plug 25 ignites the air-fuel mixture in the central portion, the SI combustion has a high combustion speed due to the high turbulent energy.

The flame of SI combustion rides on a strong swirl flow in the combustion chamber 17 and propagates in the circumferential direction. In the CI combustion, CI combustion is performed from the outer peripheral portion to the central portion in the combustion chamber 17.

When a strong swirl flow is generated in the combustion chamber 17, SI combustion can be sufficiently performed by the start of CI combustion. It is possible to suppress the generation of combustion noise and the variation in torque between cycles.

(Engine operating area)
6 and 7 illustrate a map relating to the control of the engine 1. The map is stored in the memory 102 of the ECU 10. The map includes three types of maps 501, 502, and 503. The ECU 10 uses a map selected from three types of maps 501, 502, and 503 for controlling the engine 1 according to the height of the wall temperature of the combustion chamber 17 and the temperature of the intake air. Details of the selection of the three types of maps 501, 502, and 503 will be described later.

The first map 501 is a warm map of the engine 1. The second map 502 is a map at the time of semi-warm-up of the engine 1. The third map 503 is a cold map of the engine 1.

Each map 501, 502, 503 is defined by the load and the rotation speed of the engine 1. The first map 501 is roughly divided into three regions according to the height of the load and the height of the rotation speed. Specifically, the three regions include a low load region A1 that includes idle operation and extends to low rotation and medium rotation regions, and medium and high load regions A2, A3, A4, and low that have a higher load than the low load region A1. It is a high rotation region A5 having a higher rotation speed than the load region A1, the medium and high load regions A2, A3, and A4. The medium-high load regions A2, A3, and A4 also have a medium-load region A2, a high-load medium-rotation region A3 having a higher load than the medium-load region A2, and a high-load low-rotation region having a lower rotation speed than the high-load medium-rotation region A3. It is divided into area A4.

The second map 502 is roughly divided into two areas. Specifically, the two regions are the low and medium rotation regions B1, B2, B3, and the high rotation region B4 having a higher rotation speed than the low and medium rotation regions B1, B2, and B3. The low / medium rotation regions B1, B2, and B3 are also divided into a low / medium load region B1 corresponding to the low load region A1 and the medium load region A2, a high load / medium rotation region B2, and a high load / low rotation region B3.

The third map 503 is not divided into a plurality of regions, and has only one region C1.

Here, in the low rotation region, the medium rotation region, and the high rotation region, when the entire operation region of the engine 1 is divided into substantially three equal parts of the low rotation region, the medium rotation region, and the high rotation region in the rotation speed direction, respectively. It may be a low rotation region, a medium rotation region, and a high rotation region. In the examples of FIGS. 6 and 7, a rotation speed of less than N1 is defined as a low rotation speed, a rotation speed of N2 or more is defined as a high rotation speed, and a rotation speed of N1 or more and less than N2 is defined as a medium rotation speed. The rotation speed N1 may be, for example, about 1200 rpm, and the rotation speed N2 may be, for example, about 4000 rpm.

Further, the low load region may be a region including a light load operating state, the high load region may be a region including a fully open load operating state, and the medium load may be a region between the low load region and the high load region. Further, in the low load region, the medium load region, and the high load region, when the entire operating region of the engine 1 is divided into substantially three equal parts of the low load region, the medium load region, and the high load region in the load direction, respectively. It may be a low load region, a medium load region, and a high load region.

Maps 501, 502, and 503 of FIG. 6 show the state of the air-fuel mixture and the combustion mode in each region, respectively. Map 504 of FIG. 7 corresponds to the first map 501, and in the map, the state and combustion mode of the air-fuel mixture in each region, the opening degree of the swirl control valve 56 in each region, and the drive region of the turbocharger 44. And the non-driving region. The engine 1 has a low load region A1, a medium load region A2, a high load medium rotation region A3, a high load low rotation region A4, a low medium load region B1, a high load medium rotation region B2, and a high load low. SPCCI combustion is performed in the rotation region B3. The engine 1 also performs SI combustion in other regions, specifically, the high rotation region A5, the high rotation region B4, and the region C1.

(Engine operation in each area)
Hereinafter, the operation of the engine 1 in each region of the map 504 of FIG. 7 will be described in detail with reference to the fuel injection timing and the ignition timing shown in FIG. The horizontal axis of FIG. 8 is the crank angle. The reference numerals 601 and 602, 603, 604, 605 and 606 in FIG. 8 correspond to the operating states of the engine 1 indicated by the reference numerals 601, 602, 603, 604, 605 and 606 in the map 504 of FIG. 7, respectively.

(Low load area)
When the engine 1 is operating in the low load region A1, the engine 1 performs SPCCI combustion.

Reference numeral 601 in FIG. 8 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 low load region A1, and the combustion waveform (that is, reference numeral 6013). A waveform showing the change in the heat generation rate with respect to the crank angle, reference numeral 6014) is shown. Reference numeral 602 indicates a fuel injection timing (reference numeral 6021, 6022) and an ignition timing (reference numeral 6023), and a combustion waveform (reference numeral 6024) when the engine 1 is operating in the operating state 602 in the low load region A1. , Reference numeral 603 indicates the fuel injection timing (reference numeral 6031, 6032) and 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 low load region A1. Shows. In the operating states 601, 602, and 603, the rotation speed of the engine 1 is the same and the load is different. The operating state 601 has the lowest load (that is, light load), the operating state 602 has the next lowest load (that is, the low load), and the operating state 603 has the highest load.

In order to improve the fuel efficiency of the engine 1, the EGR system 55 introduces EGR gas into the combustion chamber 17. Specifically, the intake electric S-VT23 and the exhaust electric S-VT24 provide a positive overlap period in which both the intake valve 21 and the exhaust valve 22 are opened near the exhaust top dead center. A part 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. Since the hot exhaust gas is introduced into the combustion chamber 17, the temperature inside the combustion chamber 17 becomes high. It is advantageous for stabilizing SPCCI combustion. The intake electric S-VT23 and the exhaust electric S-VT24 may be provided with a negative overlap period in which both the intake valve 21 and the exhaust valve 22 are closed.

Further, the swirl generating portion forms a strong swirl flow in the combustion chamber 17. The swirl ratio is, for example, 4 or more. The swirl control valve 56 has a predetermined opening degree on the fully closed or closed 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 multiple times during the intake stroke (reference numerals 6011, 6012, 6021, 6022, 6031, 6032). The air-fuel mixture is stratified by a plurality of fuel injections and a swirl flow in the combustion chamber 17.

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 A / F of the air-fuel mixture in the central portion is 20 or more and 30 or less, and the 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 is the same in 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, by setting the A / F of the air-fuel mixture on the outer peripheral portion to 35 or more, CI combustion is stabilized.

The air-fuel ratio (A / F) of the air-fuel mixture is leaner than the theoretical air-fuel ratio in the entire combustion chamber 17 (that is, the excess air ratio λ> 1). More specifically, the A / F of the air-fuel mixture is 30 or more in the entire combustion chamber 17. By doing so, it is possible to suppress the generation of RawNOx and improve the exhaust gas performance.

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

When the load of the engine 1 is high in the operating state 602, the injector 6 starts the second injection 6022 performed in the latter 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 amount ratio between the first injection 6021 and the second injection 6022 may be, for example, 7: 3 to 8: 2.

When the load of the engine 1 is higher in the operating state 603, the injector 6 starts the second injection 6032 performed in the latter 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 more homogeneous. The injection amount ratio between the first injection 6031 and the second injection 6032 may be, for example, 6: 4.

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

As described above, since the fuel concentration of the air-fuel mixture in the central portion is relatively high, 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. In SPCCI combustion, the controllability of CI combustion is improved. The generation of combustion noise is suppressed. Further, by making the A / F of the air-fuel mixture leaner than the stoichiometric air-fuel ratio and performing SPCCI combustion, the fuel efficiency performance of the engine 1 can be significantly improved. The low load region A1 corresponds to layer 3 described later. Layer 3 extends to the light load operating region and includes the minimum load operating state.

(Medium and high load area)
Even when the engine 1 is operating in the medium-high load region, the engine 1 performs SPCCI combustion as in the low-load region.

Reference numeral 604 in FIG. 8 indicates the fuel injection timing (reference numeral 6041, 6042) and the ignition timing (reference numeral 6043) when the engine 1 is operating in the operating state 604 in the medium load region A2 even in the medium / high load region. , Combustion waveform (reference numeral 6044) is shown. Reference numeral 605 indicates a fuel injection timing (reference numeral 6051), an ignition timing (reference numeral 6052), and a combustion waveform (reference numeral 6053) when the engine 1 is operating in the operating state 605 in the high load low rotation region A4. ing.

The EGR system 55 introduces EGR gas into the combustion chamber 17. Specifically, the intake electric S-VT23 and the exhaust electric S-VT24 provide a positive overlap period in which both the intake valve 21 and the exhaust valve 22 are opened near the exhaust top dead center. The internal EGR gas is introduced into the combustion chamber 17. Further, 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. The external EGR gas adjusts the temperature inside the combustion chamber 17 to an appropriate 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 the EGR gas, including the internal EGR gas and the external EGR gas, at full open load.

Further, in the medium-high load region A2 and the high-load medium rotation region A3, the swirl control valve 56 has a predetermined opening degree on the fully closed or closed side. A strong swirl flow having a swirl ratio of 4 or more is formed in the combustion chamber 17. On the other hand, in the high load low rotation region A4, the swirl control valve 56 is open.

The air-fuel ratio (A / F) of the air-fuel mixture is the theoretical air-fuel ratio (A / F≈14.7) in the entire combustion chamber 17. The three-way catalyst 511 and 513 purify the exhaust gas discharged from the combustion chamber 17, so that the exhaust gas performance of the engine 1 is improved. The A / F of the air-fuel mixture may be contained in the purification window of the three-way catalyst. The excess air ratio λ of the air-fuel mixture may be 1.0 ± 0.2. When the engine 1 is operating in the high load medium rotation region A3 including the fully open load (that is, the maximum load), the A / F of the air-fuel mixture is the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio in the entire combustion chamber 17. It may be richer than (that is, the air-fuel ratio λ of the air-fuel mixture is λ ≦ 1).

Since the EGR gas is introduced into the combustion chamber 17, the G / F, which is the weight ratio of the total gas in the combustion chamber 17 to the fuel, is 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 at 18 or more and 30 or less. Further, the G / F may be set at 18 or more and 50 or less.

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

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

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

By performing SPCCI combustion with the A / F of the air-fuel mixture as the stoichiometric air-fuel ratio, the exhaust gas discharged from the combustion chamber 17 can be purified by utilizing the three-way catalysts 511 and 513. Further, by introducing the EGR gas into the combustion chamber 17 to dilute the air-fuel mixture, the fuel efficiency performance of the engine 1 is improved. The medium and high load regions A2, A3, and A4 correspond to layer 2 described later. Layer 2 extends to the high load region and includes the maximum load operating state.

(Operation of turbocharger)
Here, as shown in the map 504 of FIG. 7, the turbocharger 44 is turned off in a part of the low load region A1 and a part of the medium-high load region A2 (see S / C OFF). Specifically, the turbocharger 44 is off in the low rotation side region in the low load region A1. In the region on the high rotation speed side in the low load region A1, the supercharger 44 is turned on in order to secure the required intake air filling amount in response to the increase in the rotation speed of the engine 1. Further, the turbocharger 44 is off in a part of the low load low rotation side in the medium / high load region A2. In the region on the high load side in the medium-high load region A2, the supercharger 44 is turned on in order to secure the required intake air filling amount in response to the increase in the fuel injection amount. Further, the turbocharger 44 is also on in the region on the high rotation side in the medium-high load region A2.

In each region of the high load medium rotation region A3, the high load low rotation region A4, and the high rotation region A5, the turbocharger 44 is turned on in the entire area (see S / C ON).

(High rotation area)
When the rotation speed of the engine 1 is high, the time required for the crank angle to change by 1 ° becomes short. It becomes difficult to stratify the air-fuel mixture in the combustion chamber 17. When the rotation speed of the engine 1 becomes high, it becomes difficult to perform SPCCI combustion.

Therefore, when the engine 1 is operating in the high rotation speed region A5, the engine 1 performs SI combustion instead of SPCCI combustion. The high rotation region A5 extends over the entire load direction from low load to high load.

Reference numeral 606 in FIG. 8 indicates the fuel injection timing (reference numeral 6061) and the ignition timing (reference numeral 6062) when the engine 1 is operating in the high load operating state 606 in the high rotation region A5, and the combustion waveform (reference numeral 6062). 6063) is shown.

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

The swirl control valve 56 is fully open. No swirl flow is generated in the combustion chamber 17, but only a tumble flow is generated. By fully opening the swirl control valve 56, the filling efficiency can be increased and the pump loss can be reduced.

The air-fuel ratio (A / F) of the air-fuel mixture is basically the theoretical 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 1.0 ± 0.2. When the engine 1 is operating in the vicinity of the fully open load, the excess air ratio λ of the air-fuel mixture may be less than 1.

The injector 6 starts fuel injection during the intake stroke. The injector 6 injects fuel all at once (reference numeral 6061). By initiating fuel injection during the intake stroke, a homogeneous or substantially homogeneous mixture is formed in the combustion chamber 17. Further, since the vaporization time of the fuel can be secured for a long time, the unburned loss can be reduced.

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

(Map layer structure)
Maps 501, 502, and 503 of the engine 1 shown in FIG. 6 are composed of a combination of three layers, layer 1, layer 2, and layer 3, as shown in FIG.

Layer 1 is a base layer. Layer 1 extends over the entire operating area of engine 1. Layer 1 corresponds to the entire third map 503.

Layer 2 is a layer that overlaps layer 1. Layer 2 corresponds to a part of the operating area of the engine 1. Specifically, layer 2 corresponds to the low and medium rotation regions B1, B2, and B3 of the second map 502.

Layer 3 is a layer that overlaps layer 2. Layer 3 corresponds to the low load region A1 of the first map 501.

Layers 1, 2 and 3 are selected according to the height of the wall temperature of the combustion chamber 17 and the temperature of the intake air.

When the wall temperature of the combustion chamber 17 is equal to or higher than the first predetermined wall temperature (for example, 80 ° C.) and the intake air temperature is equal to or higher than the first predetermined intake air temperature (for example, 50 ° C.), layer 1, layer 2 and layer 3 are selected. The first map 501 is configured by superimposing the layers 1, the layer 2 and the layer 3. In the low load region A1 in the first map 501, the highest layer 3 is enabled there, and in the medium and high load regions A2, A3, A4, the highest layer 2 is enabled there, and the high rotation region A5 is. , Layer 1 is enabled.

The wall temperature of the combustion chamber 17 is less than the first predetermined wall temperature, the second predetermined wall temperature (for example, 30 ° C.) or more, and the intake air temperature is less than the first predetermined intake air temperature, the second predetermined intake air temperature (for example, 25 ° C.) or more. Occasionally, layer 1 and layer 2 are selected. The second map 502 is formed by superimposing these layers 1 and 2 on top of each other. In the low and medium rotation regions B1, B2, and B3 in the second map 502, the uppermost layer 2 is valid, and in the high rotation region B4, layer 1 is valid.

When the wall temperature of the combustion chamber 17 is lower than the second predetermined wall temperature and the intake air temperature is lower than the second predetermined intake air temperature, only layer 1 is selected and the third map 503 is configured.

The wall temperature of the combustion chamber 17 may be substituted by, for example, the temperature of the cooling water of the engine 1 measured by the water temperature sensor SW10. Further, the wall temperature of the combustion chamber 17 may be estimated based on the temperature of the cooling water and other measurement results. Further, the intake air temperature can be measured by, for example, the third intake air temperature sensor SW17 that measures the temperature inside the surge tank 42. Further, the intake air temperature introduced into the combustion chamber 17 may be estimated based on various measurement results.

As described above, SPCCI combustion is performed by generating a strong swirl flow in the combustion chamber 17. Since the flame propagates along the wall of the combustion chamber 17 in SI combustion, the flame propagation of SI combustion is affected by the wall temperature. If the wall temperature is low, the flame of SI combustion will be cooled and the timing of compression ignition will be delayed.

Since CI combustion in SPCCI combustion is performed from the outer peripheral portion to the central portion of the combustion chamber 17, it is affected by the temperature of the central portion of the combustion chamber 17. If the temperature in the central part is low, CI combustion becomes unstable. The temperature of the central portion of the combustion chamber 17 depends on the temperature of the intake air introduced into the combustion chamber 17. That is, when the intake air temperature is high, the temperature in the central portion of the combustion chamber 17 is high, and when the intake air temperature is low, the temperature in the central portion is low.

When the wall temperature of the combustion chamber 17 is lower than the second predetermined wall temperature and the intake air temperature is lower than the second predetermined intake air temperature, SPCCI combustion cannot be stably performed. Therefore, only layer 1 that executes SI combustion is selected, and the ECU 10 operates the engine 1 based on the third map 503. Combustion stability can be ensured by the engine 1 performing SI combustion in all operating regions.

When the wall temperature of the combustion chamber 17 is equal to or higher than the second predetermined wall temperature and the intake air temperature is equal to or higher than the second predetermined intake temperature, the air-fuel mixture having a substantially stoichiometric air-fuel ratio (that is, λ≈1) can be stably burned by SPCCI. can. Therefore, in addition to layer 1, layer 2 is selected, and the ECU 10 operates the engine 1 based on the second map 502. When the engine 1 performs SPCCI combustion in a part of the operating region, the fuel efficiency performance of the engine 1 is improved.

When the wall temperature of the combustion chamber 17 is equal to or higher than the first predetermined wall temperature and the intake air temperature is equal to or higher than the first predetermined intake temperature, the air-fuel mixture leaner than the stoichiometric air-fuel ratio can be stably burned by SPCCI. Therefore, in addition to layer 1 and layer 2, layer 3 is selected, and the ECU 10 operates the engine 1 based on the first map 501. The fuel efficiency performance of the engine 1 is further improved by the engine 1 burning the lean air-fuel mixture in SPCCI in a part of the operating region.

Next, a control example related to the layer selection of the map executed by the ECU 10 will be described with reference to the flowchart of FIG. First, in step S1 after the start, the ECU 10 reads the signals of the sensors SW1 to SW17. In the subsequent step S2, the ECU 10 determines whether or not the wall temperature of the combustion chamber 17 is 30 ° C. or higher and the intake air temperature is 25 ° C. or higher. If the determination in step S2 is YES, the process proceeds to step S3, and if NO, the process proceeds to step S5. The ECU 10 selects only layer 1 in step S5. The ECU 10 operates the engine 1 based on the third map 503. The process then returns.

In step S3, the ECU 10 determines whether or not the wall temperature of the combustion chamber 17 is 80 ° C. or higher and the intake air temperature is 50 ° C. or higher. If the determination in step S3 is YES, the process proceeds to step S4, and if NO, the process proceeds to step S6.

The ECU 10 selects layer 1 and layer 2 in step S6. The ECU 10 operates the engine 1 based on the second map 502. The process then returns.

The ECU 10 selects layer 1, layer 2, and layer 3 in step S4. The ECU 10 operates the engine 1 based on the first map. The process then returns.

(Valve timing of intake valve and exhaust valve)
FIG. 11 shows an example of a change in the valve opening timing IVO of the intake valve 21 when the ECU 10 controls the intake electric S-VT23 according to the control logic set for the layer 2. The upper view of FIG. 11 (that is, graph 1101) shows the change (horizontal axis) of the valve opening timing IVO of the intake valve 21 with respect to the height (horizontal axis) of the load of the engine 1. The solid line corresponds to the first rotation speed when the rotation speed of the engine 1 is relatively low, and the broken line corresponds to the second rotation speed when the rotation speed of the engine 1 is relatively high (first rotation speed <second rotation speed). do.

The lower figure of FIG. 11 (that is, graph 1102) shows the change (vertical axis) of the valve opening timing IVO of the intake valve 21 with respect to the high / low (horizontal axis) of the rotation speed of the engine 1. The solid line corresponds to the first load when the load of the engine 1 is relatively low, and the broken line corresponds to the second load when the load of the engine 1 is relatively high (first load <second load).

In the graph 1101 and the graph 1102, the valve opening timing IVO of the intake valve 21 advances as it goes up, and the positive overlap period in which both the intake valve 21 and the exhaust valve 22 open becomes longer. Therefore, the amount of EGR gas introduced into the combustion chamber 17 increases.

In layer 2, the engine 1 operates with the A / F of the air-fuel mixture set to the stoichiometric air-fuel ratio or the substantially stoichiometric air-fuel ratio, and the G / F set to leaner than the stoichiometric air-fuel ratio. When the load of the engine 1 is low, the fuel supply amount is small. As shown in the graph 1101, when the load of the engine 1 is low, the ECU 10 sets the valve opening timing IVO of the intake valve 21 to the timing on the retard side. The amount of EGR gas introduced into the combustion chamber 17 is limited, and combustion stability is ensured.

When the load of the engine 1 becomes high, the fuel supply amount increases, so that the combustion stability is improved. The ECU 10 sets the valve opening timing of the intake valve 21 to the timing on the advance angle side. By increasing the amount of EGR gas introduced into the combustion chamber 17, the pump loss of the engine 1 can be reduced.

When the load of the engine 1 becomes higher, the temperature in the combustion chamber 17 becomes higher. The amount of internal EGR gas is reduced and the amount of external EGR gas is increased so that the temperature inside the combustion chamber 17 does not become too high. Therefore, the ECU 10 sets the valve opening timing of the intake valve 21 to the timing on the retard side again.

When the load of the engine 1 becomes higher and the supercharger 44 supercharges, the ECU 10 sets the valve opening timing of the intake valve 21 to the timing on the advance side again. Since a positive overlap period is provided in which both the intake valve 21 and the exhaust valve 22 are opened, the residual gas in the combustion chamber 17 can be scavenged.

It should be noted that the tendency of the valve opening timing of the intake valve 21 to change is substantially the same when the rotation speed of the engine 1 is high and when the rotation speed is low.

As shown in Graph 1102, when the rotation speed of the engine 1 is low, the flow in the combustion chamber 17 becomes weak. Since the combustion stability is lowered, the amount of EGR gas introduced into the combustion chamber 17 is limited. The ECU 10 sets the valve opening timing of the intake valve 21 to the timing on the retard side.

As the rotation speed of the engine 1 increases, the flow in the combustion chamber 17 becomes stronger, so that the amount of EGR gas introduced into the combustion chamber 17 can be increased. The ECU 10 sets the valve opening timing of the intake valve 21 to the timing on the advance angle side.

When the rotation speed of the engine 1 becomes higher, the ECU 10 sets the valve opening timing of the intake valve 21 to the timing on the retard side corresponding to the rotation speed of the engine 1. This maximizes the amount of gas introduced into the combustion chamber 17.

FIG. 12 shows the valve opening timing IVO of the intake valve 21 and the valve closing timing EVC of the exhaust valve 22 when the ECU 10 controls the intake electric S-VT23 and the exhaust electric S-VT24 according to the control logic set for the layer 3. , And, an example of the change of the overlap period O / L of the intake valve 21 and the exhaust valve 22 is shown.

The upper view of FIG. 12 (that is, graph 1201) shows the change (horizontal axis) of the valve opening timing IVO of the intake valve 21 with respect to the high / low load (horizontal axis) of the engine 1. The solid line corresponds to the third rotation speed when the engine speed is relatively low, and the broken line corresponds to the fourth rotation speed (third rotation speed <fourth rotation speed) when the rotation speed of the engine 1 is relatively high. do.

The middle figure of FIG. 12 (that is, graph 1202) shows the change (horizontal axis) of the valve closing timing EVC of the exhaust valve 22 with respect to the height (horizontal axis) of the load of the engine 1. The solid line corresponds to the case where the rotation speed of the engine 1 is the third rotation speed, and the broken line corresponds to the case where the rotation speed of the engine 1 is the fourth rotation speed.

The lower figure of FIG. 12 (that is, graph 1203) shows the change (horizontal axis) of the overlap period O / L between the intake valve 21 and the exhaust valve 22 with respect to the height (horizontal axis) of the load of the engine 1. The solid line corresponds to the case where the rotation speed of the engine 1 is the third rotation speed, and the broken line corresponds to the case where the rotation speed of the engine 1 is the fourth rotation speed.

In layer 3, the engine 1 is operated by SPCCI combustion of an air-fuel mixture whose A / F is leaner than the stoichiometric air-fuel ratio. When the load of the engine 1 is low, the fuel supply amount is small. The ECU 10 limits the amount of gas introduced into the combustion chamber 17 so that the A / F of the air-fuel mixture does not become too low. As shown in graph 1201, the ECU 10 sets the valve opening timing IVO of the intake valve 21 to the timing on the retard side of the exhaust top dead center. The valve closing timing of the intake valve 21 is the so-called late closing after the intake bottom dead center. Further, when the load of the engine 1 is low, the ECU 10 limits the amount of internal EGR gas introduced into the combustion chamber 17. As shown in graph 1202, the ECU 10 sets the valve closing timing EVC of the exhaust valve 22 to the timing on the advance angle side. The valve closing timing EVC of the exhaust valve 22 approaches the exhaust top dead center.

Since the fuel supply amount increases as the load of the engine 1 increases, the ECU 10 does not limit the amount of gas introduced into the combustion chamber 17. Further, in order to stabilize the SPCCI combustion of the air-fuel mixture leaner than the stoichiometric air-fuel ratio, the ECU 10 increases the amount of the internal EGR gas introduced into the combustion chamber 17. The ECU 10 sets the valve opening timing IVO of the intake valve 21 at a timing on the advance angle side of the exhaust top dead center. Further, the ECU 10 sets the valve closing timing EVC of the exhaust valve 22 at a timing on the retard side of the exhaust top dead center. As a result, as shown in Graph 1203, when the load of the engine 1 becomes high, the overlap period in which both the intake valve 21 and the exhaust valve 22 are opened becomes long.

When the load of the engine 1 becomes higher, the ECU 10 reduces the amount of the internal EGR gas introduced into the combustion chamber 17 so that the temperature in the combustion chamber 17 does not become too high. The ECU 10 brings the valve closing timing EVC of the exhaust valve 22 closer to the exhaust top dead center. The overlap period between the intake valve 21 and the exhaust valve 22 is shortened. Further, when the load of the engine 1 is high and the rotation speed of the engine 1 is high, the ECU 10 sets the valve opening timing of the intake valve 21 to the retard side timing as compared with the case where the rotation speed is low. The amount of gas introduced into the combustion chamber 17 is maximized.

In the low load region surrounded by the alternate long and short dash line in FIG. 12, the engine 1 may be operated in a reduced cylinder for the purpose of improving fuel efficiency. When the cylinder reduction operation is performed, the amount of gas introduced into the combustion chamber 17 and the amount of internal EGR gas are not limited. As shown by the two-dot chain line in graphs 1201 and 1202, the ECU 10 sets the valve opening timing of the intake valve 21 to the timing on the advance angle side, and sets the valve closing timing of the exhaust valve 22 to the timing on the retard angle side. It may be set.

(Engine control logic)
FIG. 13 is a flowchart showing the control logic of the engine 1. The ECU 10 operates the engine 1 according to the control logic stored in the memory 102. Specifically, the ECU 10 determines the operating state of the engine 1 based on the signals of the sensors SW1 to SW17, and burns the combustion chamber 17 so that the combustion has an SI rate according to the operating state. Calculations for adjusting the state amount in the chamber 17, adjusting the injection amount, adjusting the injection timing, and adjusting the ignition timing are performed.

First, the ECU 10 reads the signals of the sensors SW1 to SW17 in step S131. Next, in step S132, the ECU 10 determines the operating state of the engine 1 based on the signals of the sensors SW1 to SW17, and sets the target SI rate (that is, the target calorific value ratio). The target SI rate is set according to the operating state of the engine 1.

FIG. 14 schematically shows an example of setting the target SI rate. The ECU 10 sets the target SI rate low when the load of the engine 1 is low, and sets the target SI rate high when the load of the engine 1 is high. When the load of the engine 1 is low, by increasing the ratio of CI combustion in SPCCI combustion, both suppression of combustion noise and improvement of fuel efficiency are achieved. When the load of the engine 1 is high, increasing the ratio of SI combustion in SPCCI combustion is advantageous in suppressing combustion noise.

Returning to the flow of FIG. 13, in the following step S133, the ECU 10 sets the target in-cylinder state amount for realizing the set target SI rate based on the combustion model stored in the memory 102. Specifically, the ECU 10 sets a target temperature and a target pressure in the combustion chamber 17, and a target state quantity. In step S134, the ECU 10 has the opening degree of the EGR valve 54, the opening degree of the throttle valve 43, the opening degree of the air bypass valve 48, the opening degree of the swirl control valve 56, which are necessary for realizing the target in-cylinder state amount. Further, the phase angles of the intake electric S-VT23 and the exhaust electric S-VT24 (that is, the valve timing of the intake valve 21 and the valve timing of the exhaust valve 22) are set. The ECU 10 sets the control amount of these devices based on the map stored in the memory 102. The ECU 10 outputs signals to the EGR valve 54, the throttle valve 43, the air bypass valve 48, the swirl control valve (SCV) 56, and the intake electric S-VT23 and the exhaust electric S-VT24 based on the set control amount. .. By operating each device based on the signal of the ECU 10, the state quantity in the combustion chamber 17 becomes the target state quantity.

The ECU 10 further calculates a predicted value and an estimated value of the state amount in the combustion chamber 17 based on the set control amount of each device. The state quantity predicted value is a value that predicts the state quantity in the combustion chamber 17 before the intake valve 21 is closed. The state quantity predicted value is used for setting the fuel injection amount in the intake stroke, as will be described later. The state quantity estimated value is a value estimated for the state quantity in the combustion chamber 17 after the intake valve 21 is closed. The state quantity estimated value is used for setting the fuel injection amount and the ignition timing in the compression stroke, as will be described later. The state quantity estimate is also used to calculate the state quantity error by comparison with the actual combustion state.

In step S135, the ECU 10 sets the fuel injection amount during the intake stroke based on the state quantity predicted value. When performing split injection during the intake stroke, the injection amount of each injection is set. When the fuel is not injected during the intake stroke, the fuel injection amount is zero. In step S136, the ECU 10 outputs a signal to the injector 6 so that the injector 6 injects fuel into the combustion chamber 17 at a predetermined injection timing.

In step S137, the ECU 10 sets the fuel injection amount during the compression stroke based on the state quantity estimated value and the fuel injection result during the intake stroke. When the fuel is not injected during the compression stroke, the fuel injection amount is zero. In step S138, the ECU 10 outputs a signal to the injector 6 so that the injector 6 injects fuel into the combustion chamber 17 at the injection timing based on the preset map.

In step S139, the ECU 10 sets the ignition timing based on the estimated state quantity and the fuel injection result during the compression stroke. The ECU 10 outputs a signal to the spark plug 25 so that the spark plug 25 ignites the air-fuel mixture in the combustion chamber 17 at the ignition timing set in step S1310.

When the spark plug 25 ignites the air-fuel mixture, SI combustion or SPCCI combustion is performed in the combustion chamber 17. In step S1311, the ECU 10 reads the change in pressure in the combustion chamber 17 measured by the acupressure sensor SW6, and determines the combustion state of the air-fuel mixture in the combustion chamber 17 based on the change. In step S1312, the ECU 10 also compares the measurement result of the combustion state with the state quantity estimated value estimated in step S134, and calculates an error between the state quantity estimated value and the actual state quantity. The calculated error is used for the estimation of step S134 in the cycle after this time. The ECU 10 has an opening degree of the throttle valve 43, the EGR valve 54, the swirl control valve 56, and / or the air bypass valve 48, and the intake electric S-VT23 and the exhaust electric S-VT24 so that the state quantity error is eliminated. Adjust the phase angle. Thereby, the amount of fresh air and EGR gas introduced into the combustion chamber 17 is adjusted.

In step S138, the ECU 10 also injects during the compression stroke so that the ignition timing can be advanced when the temperature in the combustion chamber 17 is predicted to be lower than the target temperature based on the estimated state quantity. Advance the timing over the map-based injection timing. On the other hand, in step S138, the ECU 10 is in the compression stroke so that the ignition timing can be retarded when the temperature in the combustion chamber 17 is predicted to be higher than the target temperature based on the estimated state quantity. The injection timing is retarded from the injection timing based on the map.

That is, when the temperature in the combustion chamber 17 is low, the timing θ _CI (see FIG. 5) at which the unburned air-fuel mixture self-ignites after the SI combustion is started by spark ignition is delayed, and the SI rate is the target. It deviates from the SI rate. In this case, the amount of unburned fuel increases and the exhaust gas performance deteriorates.

Therefore, when it is predicted that the temperature in the combustion chamber 17 will be lower than the target temperature, the ECU 10 advances the injection timing and advances the ignition timing in step S1310. Since sufficient heat can be generated by SI combustion by accelerating the start of SI combustion, it is possible to prevent the timing θ _CI of self-ignition of the unburned air-fuel mixture from being delayed when the temperature in the combustion chamber 17 is low. be able to. As a result, the SI rate approaches the target SI rate.

Further, if the temperature in the combustion chamber 17 is high, the unburned air-fuel mixture self-ignites immediately after SI combustion is started by spark ignition, and the SI rate deviates from the target SI rate. In this case, the combustion noise increases.

Therefore, when it is predicted that the temperature in the combustion chamber 17 will be higher than the target temperature, the ECU 10 retards the injection timing and retards the ignition timing in step S1310. Since the start of SI combustion is delayed, it is possible to prevent the timing θ _CI of self-ignition of the unburned air-fuel mixture from being accelerated when the temperature in the combustion chamber 17 is high. As a result, the SI rate approaches the target SI rate.

The control logic of the engine 1 adjusts the SI rate by a state quantity setting device including a throttle valve 43, an EGR valve 54, an air bypass valve 48, a swirl control valve 56, an intake electric S-VT23, and an exhaust electric S-VT24. It is configured to do. By adjusting the state quantity in the combustion chamber 17, the SI rate can be roughly adjusted. The control logic of the engine 1 is also configured to adjust the SI rate by adjusting the fuel injection timing and the ignition timing. By adjusting the injection timing and the ignition timing, for example, it is possible to correct the difference between cylinders and finely adjust the self-ignition timing. By adjusting the SI rate in two stages, the engine 1 can accurately realize the target SPCCI combustion corresponding to the operating state.

The ECU 10 also outputs a signal to at least the EGR system 55 and the spark plug 25 so that the actual SI rate due to combustion becomes the target SI rate. Further, as described above, when the load of the engine 1 is high, the ECU 10 has a higher target SI rate than when the load is low. Therefore, when the load of the engine 1 is high, the ECU 10 has a higher SI rate than when the load is low. At least, a signal will be output to the EGR system 55 and the spark plug 25.

The control of the engine 1 performed by the ECU 10 is not limited to the control logic based on the combustion model described above.

(Design method of engine control logic)
When designing the control logic of the engine 1 that executes the above-mentioned SPCCI combustion, parameters related to the control amount of each device are set. For example, in the case of the spark plug 25, the ignition energy and the ignition timing corresponding to the operating state of the engine 1 are set. In the case of the intake electric S-VT23, the valve timing of the intake valve 21 corresponding to the operating state of the engine 1 is set. Hereinafter, the valve timing setting of the intake valve 21 in the design method of the control logic of the engine 1 will be described with reference to the drawings.

In the engine 1 that executes SPCCI combustion, in order to suppress combustion noise and realize stable SPCCI combustion, the temperature inside the combustion chamber 17 at the timing when CI combustion starts (θ _CI : see FIG. 5) is appropriate. The inventors of the present application have found that it is necessary to adjust the temperature to a high temperature. That is, if the temperature inside the combustion chamber 17 is low, the ignitability of CI combustion decreases. When the temperature inside the combustion chamber 17 is high, the combustion noise becomes large.

Timing at which CI combustion starts θ The temperature inside the combustion chamber 17 in _CI is related to the effective compression ratio of the engine 1. The effective compression ratio of the engine 1 is determined by the geometric compression ratio ε and the valve closing timing IVC of the intake valve 21. The inventors of the present application have newly found that an appropriate IVC range exists in the range of the geometric compression ratio ε in which SPCCI combustion can occur in order to put an engine capable of SPCCI combustion into practical use. The technology disclosed here requires a predetermined relationship between the geometric compression ratio ε and the valve closing timing IVC of the intake valve 21 in order to put into practical use an engine having a unique combustion mode called SPCCI combustion. There is something new in the fact that it was found to be. Further, the relational expression between the geometric compression ratio ε and the valve closing timing IVC of the intake valve 21, which will be described later, is also new.

Further, the technique disclosed herein is based on the relationship between the geometric compression ratio ε and the valve closing timing IVC of the intake valve 21 when designing the control logic of the engine 1 that performs SPCCI combustion. There is also a new point in setting the valve closing timing IVC of 21.

When the engine 1 performs SPCCI combustion by making the A / F of the air-fuel mixture richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio and making the G / F leaner than the stoichiometric air-fuel ratio (layer 2), the air-fuel mixture of the air-fuel mixture When SPCCI combustion is performed with the A / F leaner than the stoichiometric air-fuel ratio (layer 3), the switch is made. The relationship between the geometric compression ratio ε in layer 2 and the valve closing timing IVC of the intake valve 21 is different from the relationship between the geometric compression ratio ε and the valve closing timing IVC of the intake valve 21 in layer 3.

(Relationship between the geometric compression ratio in layer 2 and the valve closing timing of the intake valve)
FIG. 15 shows the characteristics of SPCCI combustion. Specifically, FIG. 15 shows the EGR ratio (EGR ratio) when the engine 1 SPCCI burns a lean air-fuel mixture with A / F having a stoichiometric air-fuel ratio and G / F having a leaner air-fuel ratio than the stoichiometric air-fuel ratio, as in layer 2. The horizontal axis) shows the range of establishment of SPCCI combustion. The vertical axis in the figure is the crank angle corresponding to the center of gravity of combustion, and the center of gravity of combustion advances as it goes up in the figure.

The range in which SPCCI combustion is established is indicated by the range marked with hatching in the figure. The range in which SPCCI combustion is established is sandwiched between the "advance limit" line and the "retard limit" line. The line of "advance angle limit" means that if the center of gravity of combustion is advanced beyond this line, abnormal combustion occurs, so that SPCCI combustion is not established. Further, the line of the "lagging limit" means that if the center of gravity of combustion is retarded from this line, self-ignition does not occur, so that SPCCI combustion is not established.

The alternate long and short dash line in the figure shows the center of gravity of combustion corresponding to MBT (Minimum advance for Best Torque). Here, the center of gravity of combustion corresponding to MBT is simply referred to as MBT. The MBT advances as the EGR rate increases.

From the viewpoint of improving fuel efficiency, it is preferable that the center of gravity of SPCCI combustion is close to that of the MBT. When the EGR rate becomes high, the range in which SPCCI combustion is established also advances, but the distance between the "advance limit" line and the "lagging limit" line becomes narrower, and the range in which SPCCI combustion is established becomes narrower.

When the load of the engine 1 is low (when the load is "light load"), the range in which SPCCI combustion is established is on the advance angle side, so the EGR rate is adjusted to a certain extent as shown by the arrows at both ends in FIG. At the same time, by adjusting the center of gravity of combustion to the advance angle side and the retard angle side, SPCCI combustion corresponding to MBT can be realized.

When the load of the engine 1 becomes high, the amount of air introduced into the combustion chamber 17 must be increased in response to the increase in the fuel supply amount. In order to increase the EGR rate with respect to the increase in the amount of air, a large amount of EGR gas must also be introduced into the combustion chamber 17. However, due to the limitation of the supercharging capacity of the turbocharger 44, it is difficult to introduce a large amount of both air and EGR gas into the combustion chamber 17. Therefore, when the load of the engine 1 becomes high, the range in which SPCCI combustion is established becomes the retard side. When the load of the engine 1 is high, if the combustion center of gravity of SPCCI combustion is to be brought closest to the MBT, SPCCI combustion must be performed at the point Y in FIG. The point Y corresponds to the operating state of the engine 1 at the upper limit load capable of SPCCI combustion of the air-fuel mixture having the A / F as the stoichiometric air-fuel ratio. When the engine 1 is operating at the upper limit load, it is difficult to realize SPCCI combustion corresponding to MBT by adjusting the EGR rate and the center of gravity of combustion.

The engine 1 operates in layer 2 over a wide operating range from low load to high load. Regarding layer 2, the state in which the engine 1 is operated with the upper limit load (maximum load) corresponds to the limit operating state in which SPCCI combustion is established. Regarding layer 2, the engine 1 is operating at the upper limit load in determining the relationship between ε and IVC so that the temperature of the combustion chamber 17 at the timing when CI combustion starts (θ _CI ) becomes a predetermined temperature. It is necessary to determine based on the temperature in the combustion chamber 17 at that time.

In order to know the temperature inside the combustion chamber 17 when the engine 1 is operating at the upper limit load, the inventors of the present application perform SPCCI combustion in the actual engine 1 and use the measured value acquired from the engine 1. I decided to do it. Specifically, the inventors of the present application measure various parameters when the engine 1 is operating at the upper limit load in layer 2, and estimate the actual temperature in the combustion chamber 17 in θ _CI from the measured parameters. did. The inventors of the present application have defined the average value of a plurality of estimated temperatures as the reference temperature Tth1. If the temperature of the combustion chamber 17 in _θCI is the reference temperature Tth1, SPCCI combustion can be realized in layer 2.

Here, in the SPCCI combustion, as described above, the θ _CI can be adjusted by adjusting the ignition timing. However, when the temperature in the combustion chamber 17 in θ _CI exceeds the reference temperature Tth1 when the engine 1 is operating at the upper limit load in layer 2, the θ _CI is adjusted even if the ignition timing is adjusted. You will not be able to. On the other hand, if the temperature in the combustion chamber 17 in θ _CI is equal to or lower than the reference temperature Tth1, the temperature in the combustion chamber 17 in θ _CI is adjusted by adjusting the ignition timing (specifically, advancing the ignition timing). Can be increased to the reference temperature Tth1 to realize SPCCI combustion.

Therefore, in order to realize SPCCI combustion in layer 2, the relationship between ε and IVC is required so that the temperature of the combustion chamber 17 in θ _CI does not exceed the reference temperature Tth1.

Therefore, as conceptually shown in FIG. 16, the inventors of the present application set the ε and the IVC in a matrix consisting of two parameters of the geometric compression ratio ε and the valve closing timing IVC of the intake valve 21. While changing the values for each, the estimation calculation of the temperature of the combustion chamber 17 at the time of θ _CI was performed using the model of the engine 1. The combination of ε and IVC having a reference temperature Tth1 or less can realize SPCCI combustion in layer 2. As shown in FIG. 16, when the geometric compression ratio ε is high and the valve closing timing IVC of the intake valve 21 approaches the intake bottom dead center, the temperature of the combustion chamber 17 during CI combustion exceeds the reference temperature Tth1. Will end up.

Note that FIG. 16 is a matrix in which the IVC is set after the inspiratory bottom dead center. Although not shown, the inventors of the present application similarly set the IVC before the dead center of intake air to the reference temperature Tth1 or less by performing the estimation calculation of the temperature of the combustion chamber 17 at the time of θ _CI . A combination of ε and IVC was obtained.

Graph 1701 in the upper figure of FIG. 17 shows approximate equations (I) and (II) calculated based on the combination of ε and IVC. The horizontal axis of the graph 1701 is the geometric compression ratio ε, and the vertical axis is the valve closing timing IVC (deg. ABDC) of the intake valve 21. Although not shown, the inventors of the present application plot the combination of ε and IVC having a reference temperature of Tth1 or less on the plane of the graph 1701, and the approximate expression (I) (I) based on the plot points. II) has been decided.

Graph 1701 corresponds to the case where the rotation speed of the engine 1 is 2000 rpm. Approximate equations (I) and (II) are, respectively.
Approximate formula (I) IVC = -0.4288ε ² + 31.518ε-379.88
Approximate formula ( ^II ) IVC = 1.9163ε 2-89.935ε ＋ 974.94
Is.

In graph 1701, in the approximate equations (I) and (II) and the combination of ε and IVC on the left side of ε = 17, the temperature of the combustion chamber 17 at the time of CI combustion becomes the reference temperature Tth1 or less. With this combination, it is possible to SPCCI burn an air-fuel mixture having an A / F having a stoichiometric air-fuel ratio and having a G / F leaner than the stoichiometric air-fuel ratio.

The relationship between ε and IVC described above is a relationship based on the upper limit temperature of the combustion chamber 17 in layer 2.

On the other hand, in layer 2, the relationship between ε and IVC must be determined so that the temperature of the combustion chamber 17 becomes a predetermined temperature even when the engine 1 is operated with a light load.

The temperature of the combustion chamber 17 that performs SPCCI combustion is the result of two pressure increases, that is, the pressure increase due to the compression work of the piston 3 in the compression stroke and the pressure increase resulting from the heat generation of SI combustion. The compression work of the piston 3 is determined by the effective compression ratio. If the effective compression ratio is too low, the pressure rise due to the compression work of the piston 3 is small. Unless the flame propagation in SPCCI combustion progresses and the pressure rise caused by the heat generated by SI combustion increases considerably, the in-cylinder temperature cannot be raised to the ignition temperature. As a result, the amount of the air-fuel mixture that is compressed and self-ignited is small, and a large amount of the air-fuel mixture is burned by flame propagation, so that the combustion period is long and the fuel efficiency is lowered. In order to stably cause CI combustion in SPCCI combustion and maximize fuel efficiency, it is necessary to keep the effective compression ratio above a certain value. Therefore, the relationship between ε and IVC must be determined.

Similar to the above, the inventors of the present application measure various parameters when the actual engine 1 is operating with a light load, and estimate the actual temperature in the combustion chamber 17 in θ _CI from the measured parameters. did. The inventors of the present application have defined the average value of a plurality of estimated temperatures as the reference temperature Tth2.

If the temperature in the combustion chamber 17 in θ _CI is equal to or higher than the reference temperature Tth2 when the engine 1 is operating with a light load, SPCCI combustion can be realized by delaying the ignition timing. However, if the temperature of the combustion chamber 17 in θ _CI is lower than the reference temperature Tth2, the temperature is too low, and even if the ignition timing is advanced, SPCCI combustion cannot be realized.

Therefore, in order to realize SPCCI combustion in layer 2, the relationship between ε and IVC is required so that the temperature of the combustion chamber 17 in θ _CI is equal to or higher than the reference temperature Tth2.

Similar to the matrix shown in FIG. 16, the inventors of the present application have values for each of ε and IVC in a matrix consisting of two parameters of the geometric compression ratio ε and the valve closing timing IVC of the intake valve 21. While changing the above, the estimation calculation of the temperature of the combustion chamber 17 at the time of CI combustion was performed using the model of the engine 1. In this matrix, the combination of ε and IVC having a reference temperature of Tth2 or higher can realize SPCCI combustion in layer 2.

Graph 1701 in FIG. 17 also shows approximate equations (III) and (IV) calculated based on the combination of ε and IVC having a reference temperature of Tth2 or higher. Approximate equations (III) and (IV) are, respectively.
Approximate formula (III) IVC = -0.4234ε ² + 22.926ε-167.84
Approximate formula (IV) IVC = 0.4234ε ^2-22.926ε ＋ 207.84
Is.

In graph 1701, in the combination of ε and IVC on the right side of the approximate equations (III) and (IV), the temperature of the combustion chamber 17 at the time of CI combustion becomes the reference temperature Tth2 or more. This combination makes it possible to SPCCI burn an air-fuel mixture having an A / F having a stoichiometric air-fuel ratio and a G / F leaner than the stoichiometric air-fuel ratio.

As can be seen from FIG. 17, the relationship between ε and IVC is substantially symmetric in the vertical direction around IVC = about 20 deg. ABDC. IVC = 20deg. ABDC corresponds to the valve closing timing (that is, the best IVC) in which the amount of gas introduced into the combustion chamber 17 is maximized when the rotation speed of the engine 1 is 2000 rpm. Further, IVC = 120deg. ABDC is the retard limit of the valve closing timing IVC of the intake valve 21, and IVC = −80deg. ABDC is the advance limit of the valve closing timing IVC of the intake valve 21.

The combination of ε and IVC within the range surrounded by the approximate equations (I), (II), (III), and (IV) in FIG. 17 is a combination that can put the engine 1 that performs SPCCI combustion into practical use in layer 2. Is. In other words, the combination of ε and IVC outside this range makes it impossible to put the engine 1 that performs SPCCI combustion in layer 2 into practical use.

The designer must determine the IVC within the shaded ε-IVC establishment range in FIG. 17 when determining the valve closing timing IVC of the intake valve 21 when the engine 1 operates in layer 2.

Specifically, the designer sets the geometric compression ratio ε to 10 ≦ ε <17.
0.4234ε 2-22.926ε ＋ 207.84 ≦ IVC ≦ －0.4234ε ^{2 ＋} 22.926ε－167.84… (1)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, when the designer sets the geometric compression ratio ε to 17 ≦ ε <20,
－0.4288ε ² ＋31.518ε－379.88 ≦ IVC ≦ －0.4234ε ² ＋22.926ε－167.84… (2)
Or,
0.4234ε 2-22.926ε ＋ 207.84 ≦ IVC ≦ 1.9163ε ^{2 －89.935ε} ＋ 974.94… (3)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, the designer sets the geometric compression ratio ε to 20 ≦ ε ≦ 30.
－0.4288ε ² ＋31.518ε－379.88 ≦ IVC ≦ 120… (4)
Or,
-80 ≤ IVC ≤ 1.9163ε ² -89.935ε + 974.94 ... (5)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

By setting the valve closing timing IVC of the intake valve 21 based on the relational expressions (1) to (5), the A / F is richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio, and the G / F is the theory. It is possible to burn SPCCI a mixture leaner than the air-fuel ratio.

The valve closing timing IVC is set for each operating state in layer 2 determined by the load and the rotation speed of the engine 1. As described above, the example shown by the solid line in FIG. 17 is the ε-IVC establishment range when the rotation speed of the engine 1 is 2000 rpm. When the number of revolutions of the engine 1 changes, the range of establishment of ε-IVC also changes. As the engine speed increases, the best IVC retards. Specifically, when the engine speed is 3000 rpm, the best IVC is about 22 deg. A BDC. As shown by the broken line in FIG. 17, the ε-IVC establishment range when the engine 1 rotation speed is 3000 rpm is only about 2 deg. Minutes with respect to the ε-IVC establishment range when the engine 1 rotation speed is 2000 rpm. , Moves in parallel to the retard side.

Therefore, when the designer sets the geometric compression ratio ε to 10 ≦ ε <17, when the rotation speed of the engine 1 is 3000 rpm, the designer sets it.
0.4234ε 2-22.926ε ＋ 209.84 ≦ IVC ≦ －0.4234ε ^{2 ＋} 22.926ε－ ^165.84 … (1-1)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, when the designer sets the geometric compression ratio ε to 17 ≦ ε <20, when the rotation speed of the engine 1 is 3000 rpm, the designer sets it.
－0.4288ε ² ＋31.518ε－377.88 ≦ IVC ≦ －0.4234ε ² ＋22.926ε－ ^165.84 … (2-1)
Or,
0.4234ε ² －22.926ε ＋ 209.84 ≦ IVC ≦ 1.9163ε ² －89.935ε ＋ ^976.94 … (3-1)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, the designer sets the geometric compression ratio ε to 20 ≦ ε ≦ 30.
-0.4288ε ² + 31.518ε- ^377.88 ≤ IVC ≤ 120 ... (4-1)
Or,
-80 ≤ IVC ≤ 1.9163ε ² -87.935ε + ^976.94 ... (5-1)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, when the rotation speed of the engine 1 is 4000 rpm, the best IVC is about 28 deg. A BDC. As shown by the alternate long and short dash line in FIG. 17, the range of establishment of ε-IVC when the rotation speed of engine 1 is 4000 rpm is about 8 deg. Minutes with respect to the range of establishment of ε-IVC when the rotation speed of engine 1 is 2000 rpm. Only move in parallel to the retard side.

Therefore, when the designer sets the geometric compression ratio ε to 10 ≦ ε <17, when the rotation speed of the engine 1 is 4000 rpm, the designer sets it.
0.4234ε 2-22.926ε ＋ 215.84 ≦ IVC ≦ －0.4234ε ² ＋ 22.926ε－ ^159.84 … ( ^1-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, when the designer sets the geometric compression ratio ε to 17 ≦ ε <20, when the rotation speed of the engine 1 is 4000 rpm, the designer sets it.
－0.4288ε ² ＋31.518ε－371.88 ≦ IVC ≦ －0.4234ε ² ＋22.926ε－159.84… (2 − ² )
Or,
0.4234ε ² －22.926ε ＋ 215.84 ≦ IVC ≦ 1.9163ε ² －89.935ε ＋ 982.94… ( ^3-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, the designer sets the geometric compression ratio ε to 20 ≦ ε ≦ 30.
－0.4288ε ² ＋31.518ε－ ^371.88 ≦ IVC ≦ 120… (4-2)
Or,
-80 ≤ IVC ≤ 1.9163ε ² -89.935ε + ^982.94 ... (5-2)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

The correction term C related to the rotation speed NE (rpm) of the engine 1 is
C = 3.3 × 10 ^-10 NE ³ －1.0 × 10 ^-6 NE ² + 7.0 × 10 ^-4 NE
Then, the relational expression between ε and IVC in layer 2 can be expressed as follows.
If the geometric compression ratio ε is 10 ≤ ε <17,
0.4234ε ^2-22.926ε ＋ 207.84 ＋ C ≦ IVC ≦ －0.4234ε ² ＋ 22.926ε －167.84 ＋ C… (1 ^-3 )
If the geometric compression ratio ε is 17 ≤ ε <20,
－0.4288ε ² ＋31.518ε－379.88 ＋ C ≦ IVC ≦ －0.4234ε ² ＋22.926ε－167.84 ＋ C… (2 ^-3 )
Or,
0.4234ε 2-22.926ε ＋ 207.84 ＋ C ≦ IVC ≦ 1.9163ε ^{2 －89.935ε} ＋ 974.94 ＋ C… (3 ^-3 )
If the geometric compression ratio ε is 20 ≤ ε ≤ 30
-0.4288ε ² + 31.518ε-379.88 + C ≤ IVC ≤ 120 ... (4 ^-3 )
Or,
-80 ≤ IVC ≤ 1.9163ε ² -89.935ε + 974.94 + C ... (5 ^-3 )
The designer determines the valve closing timing IVC based on the ε-IVC establishment range determined for each rotation speed of the engine 1. As a result, the designer can set the valve timing of the intake valve 21 in layer 2 as illustrated in FIG.

(Change in ε-IVC establishment range due to difference in octane number)
Graph 1701 in FIG. 17 shows the relationship between ε and IVC when the fuel is a high octane fuel (octane number is about 96). Graph 1702 shown in the figure below shows the relationship between ε and IVC when the fuel is a low octane number fuel (octane number is about 91). According to the study by the inventors of the present application, when the fuel has a low octane number, the ε-IVC formation range shifts toward the low compression ratio by 1.3 compression ratio with respect to the ε-IVC formation range of the high octane fuel. I understood.

Therefore, when the engine 1 with a low octane number fuel is set to the valve closing timing IVC, the designer sets the geometric compression ratio ε to 10 ≦ ε <15.7, and when the rotation speed of the engine 1 is 2000 rpm, the designer sets the geometric compression ratio ε.
0.4234ε ² −21.826ε ＋ 178.75 ≦ IVC ≦ －0.4234ε ² ＋ 21.826ε－138.75… (6)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, if the geometric compression ratio ε is set to 15.7 ≦ ε <18.7 in the engine 1 having a low octane number fuel, the designer determines that the rotation speed of the engine 1 is 2000 rpm.
－0.5603ε ² ＋34.859ε－377.22 ≦ IVC ≦ －0.4234ε ² ＋21.826ε－138.75… (7)
Or,
0.4234ε ² －21.826ε ＋ 178.75 ≦ IVC ≦ 1.9211ε ² -85.076ε ＋ 862.01… (8)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, in the engine 1 of the low octane number fuel, the designer sets the geometric compression ratio ε to 18.7 ≦ ε ≦ 30, and when the rotation speed of the engine 1 is 2000 rpm, the designer sets the geometric compression ratio ε to 18.7 ≦ ε ≦ 30.
－0.5603ε ² ＋34.859ε－377.22 ≦ IVC ≦ 120… (9)
Or,
-80 ≤ IVC ≤ 1.9211ε ^2-85.076ε + 862.01… (10)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

The range in which the cross-hatching is added to the graph 1702 in FIG. 17 is the range in which the ε-IVC formation range of the high octane fuel and the ε-IVC formation range of the low octane fuel overlap. If the IVC is defined within the range where the two establishment ranges overlap, the designer can set the control logic suitable for both the engine 1 using the high octane fuel and the engine 1 using the low octane fuel. Even if the octane number of the fuel is different for each destination, the designer can design the control logic of the engine collectively. Batch design has the advantage of reducing design man-hours.

Although not shown, even in the engine 1 with a low octane number fuel, when the rotation speed of the engine 1 increases, the ε-IVC establishment range translates to the retard side. In the engine 1 with a low octane number fuel, the designer sets the geometric compression ratio ε to 10 ≦ ε <15.7, and when the rotation speed of the engine 1 is 3000 rpm, the designer sets it.
0.4234ε ² -21.826ε + 180.75 ≤ IVC ≤ -0.4234ε ² + 21.826ε- ^136.75 ... (6-1)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, if the geometric compression ratio ε is set to 15.7 ≦ ε <18.7 in the engine 1 having a low octane number fuel, the designer determines that the rotation speed of the engine 1 is 3000 rpm.
-0.5603ε ² ＋34.859ε－375.22 ≦ IVC ≦ －0.4234ε ² ＋21.826ε－ ^136.75 … (7-1)
Or,
0.4234ε ² －21.826ε ＋ 180.75 ≦ IVC ≦ 1.9211ε ² -85.076ε ＋ 864.01… (8 ^-1 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, the designer sets the geometric compression ratio ε to 18.7 ≦ ε ≦ 30 in the engine 1 having a low octane number fuel, and when the rotation speed of the engine 1 is 3000 rpm, the designer sets the geometric compression ratio ε to 18.7 ≦ ε ≦ 30.
-0.5603ε ² +34.859ε-375.22 ≤ IVC ≤ 120 ... (9 ^-1 )
Or,
-80 ≤ IVC ≤ 1.9211ε ^2-85.076ε ＋ 864.01… (10 ^-1 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, when the geometric compression ratio ε is set to 10 ≦ ε <15.7 in the engine 1 having a low octane number fuel, the designer determines that the rotation speed of the engine 1 is 4000 rpm.
0.4234ε ² -21.826ε + 186.75 ≤ IVC ≤ -0.4234ε ² + 21.826ε-130.75 ... (6 ^-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, if the geometric compression ratio ε is set to 15.7 ≦ ε <18.7 in the engine 1 having a low octane number fuel, the designer determines that the rotation speed of the engine 1 is 4000 rpm.
-0.5603ε ² ＋34.859ε－369.22 ≦ IVC ≦ －0.4234ε ² ＋21.826ε－ ^130.75 … (7-2)
Or,
0.4234ε ² -21.826ε + 186.75 ≤ IVC ≤ 1.9211ε ^2-85.076ε + 870.01 ... (8 ^-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, the designer sets the geometric compression ratio ε to 18.7 ≦ ε ≦ 30 in the engine 1 having a low octane number fuel, and when the rotation speed of the engine 1 is 4000 rpm, the designer sets the geometric compression ratio ε to 18.7 ≦ ε ≦ 30.
-0.5603ε ² +34.859ε-369.22 ≤ IVC ≤ 120 ... (9 ^-2 )
Or,
-80 ≤ IVC ≤ 1.9211ε ² -77.076ε + 870.01 ... (10 ^-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Similarly to the above, when the correction term C relating to the rotation speed NE (rpm) of the engine 1 is used, the relational expression between ε and IVC in layer 2 in the engine 1 having a low octane number can be expressed as follows. can.
If the geometric compression ratio ε is 10 ≤ ε <15.7,
0.4234ε ² －21.826ε ＋ 178.75 ＋ C ≦ IVC ≦ －0.4234ε ² ＋ 21.826ε－138.75 ＋ C… (6 ^-3 )
If the geometric compression ratio ε is 15.7 ≤ ε <18.7,
-0.5603ε ² + 34.859ε-377.22 + C ≤ IVC ≤ -0.4234ε ² + 21.826ε-138.75 + C ... (7 ^-3 )
Or,
0.4234ε ² －21.826ε ＋ 178.75 ＋ C ≦ IVC ≦ 1.9211ε ² -85.076ε ＋ 862.01 ＋ C… (8 ^-3 )
If the geometric compression ratio ε is 18.7 ≤ ε ≤ 30
-0.5603ε ² ＋34.859ε－377.22 ＋ C ≦ IVC ≦ 120… (9 ^-3 )
Or,
-80 ≤ IVC ≤ 1.9211ε ^2-85.076ε + 862.01 + C ... (10 ^-3 )
(Relationship between the geometric compression ratio in layer 3 and the valve closing timing of the intake valve)
FIG. 18 shows the characteristics of SPCCI combustion when the engine 1 SPCCI burns an air-fuel mixture leaner than the stoichiometric air-fuel ratio in A / F as in layer 3. Specifically, FIG. 18 shows a range in which SPCCI combustion is stable with respect to G / F (horizontal axis). The vertical axis in the figure is the crank angle corresponding to the center of gravity of combustion, and the center of gravity of combustion advances as it goes up in the figure.

The range in which SPCCI combustion is stable is the range surrounded by a curve in the figure. When the load of the engine 1 is low, the range in which the SPCCI combustion is stable is located in the upper left of FIG. When the load of the engine 1 becomes high, the range in which the SPCCI combustion stabilizes moves downward in FIG.

FIG. 18 also shows a range in which the emission of RawNOx can be suppressed. The range in which the emission of RawNOx can be suppressed is located at the lower right in FIG. This range has a triangular shape in FIG. If the A / F is leaner than the stoichiometric air-fuel ratio, the RawNOx cannot be purified by the three-way catalyst. At layer 3, the engine 1 must satisfy both ensuring the stability of SPCCI combustion and suppressing the emission of RawNOx.

As can be seen from the figure, when the load of the engine 1 is high, the area where the range where the combustion stability is ensured and the range where the emission of RawNOx can be suppressed becomes large. On the other hand, when the load of the engine 1 is low, the area where the range where the combustion stability is ensured and the range where the emission of RawNOx can be suppressed becomes small.

With respect to layer 3, the state in which the engine 1 is operated with a light load corresponds to the limit operating state in which SPCCI combustion is established. Regarding layer 3, the engine 1 is operating with a light load in determining the relationship between ε and IVC so that the temperature of the combustion chamber 17 at the timing when CI combustion starts (θ _CI ) becomes a predetermined temperature. It is necessary to determine based on the temperature in the combustion chamber 17 at that time.

In the same manner as described above, the inventors of the present application measure various parameters when operating with a light load in layer 3 in the actual engine 1, and from the measured parameters, actually in the combustion chamber 17 in θ _CI . Estimated temperature. Then, the average value of the estimated plurality of temperatures was set as the reference temperature Tth3. If the temperature of the combustion chamber 17 in _θCI is the reference temperature Tth3, SPCCI combustion can be realized in layer 3. This reference temperature Tth3 corresponds to the lower limit temperature. When the engine 1 is operating with a light load, if the temperature in the combustion chamber 17 in θ _CI is equal to or higher than the reference temperature Tth3, SPCCI combustion can be realized by delaying the ignition timing. However, if the temperature in the combustion chamber 17 in θ _CI is less than the reference temperature Tth3, SPCCI combustion cannot be realized even if the ignition timing is advanced.

Therefore, in order to realize stable SPCCI combustion in layer 3, the relationship between ε and IVC is required so that the temperature in the combustion chamber 17 in θ _CI is equal to or higher than the reference temperature Tth3.

Therefore, as conceptually shown in FIG. 19, the inventors of the present application have set the valve closing timing IVC of the intake valve 21 and the intake valve 21 and the intake valve 21 for each geometric compression ratio ε (ε1, ε2, ...). Estimating the temperature of the combustion chamber 17 during CI combustion using the model of engine 1 while changing the values for IVC and O / L in the matrix with the overlap period O / L of the exhaust valve 22. (Reference: 1901). As shown by hatching the matrix of reference numeral 1901, the combination of IVC and O / L having a reference temperature Tth3 or higher can realize SPCCI combustion in layer 3.

Further, in order to suppress the discharge of RawNOx, the G / F of the air-fuel mixture must be a predetermined value or more. As shown by reference numeral 1902 in FIG. 19, the inventors of the present application changed the values for IVC and O / L in the matrix consisting of the two parameters of the valve closing timing IVC and the overlap period O / L. However, the G / F estimation calculation was performed using the model of the engine 1. As shown by the shaded lines in the matrix of reference numeral 1902, the combination of IVC and O / L having a G / F of a predetermined value or more can suppress the emission of RawNOx.

Then, the inventors of the present application have a combination of IVC and O / L whose reference temperature is Tth3 or higher, which is indicated by reference numeral 1901, and IVC and O / L, which are indicated by reference numeral 1902 and whose G / F is equal to or higher than a predetermined value. By superimposing the combination, the relationship between ε and IVC, which can realize both the stability of SPCCI combustion and the suppression of RawNOx emission, was determined. That is, in the matrix of reference numeral 1903, the cross-hatched range is a combination of ε and IVC that can achieve both stability of SPCCI combustion and suppression of NOx emissions.

Although not shown, the inventors of the present application similarly perform an estimation calculation of the temperature of the combustion chamber 17 during CI combustion and G for a matrix in which the valve closing timing of the intake valve 21 is set before the intake bottom dead center. By performing the estimation calculation of / F, a combination of IVC and O / L having a reference temperature of Tth3 or higher and a combination of IVC and O / L having a G / F of predetermined temperature or higher were obtained.

FIG. 20 shows approximate equations (V) and (VI) calculated from the combination of ε and IVC. The horizontal axis of FIG. 20 is the geometric compression ratio ε, and the vertical axis is the valve closing timing IVC (deg. ABDC) of the intake valve 21.

The upper figure 2001 of FIG. 20 corresponds to the case where the rotation speed of the engine 1 is 2000 rpm. Approximate equations (V) and (VI) are, respectively.
Approximate formula (V) IVC = -0.9949ε ² + 41.736ε-361.16
Approximate formula (VI) IVC = 0.9949ε ^2-41.736ε + 401.16
Is.

In FIG. 20, in the combination of ε and IVC on the right side of the approximate equations (V) and (VI), the temperature of the combustion chamber 17 during CI combustion is the reference temperature Tth3 or higher, and the A / F is higher than the stoichiometric air-fuel ratio. SPCCI combustion of lean air-fuel mixture is realized.

The relationship between ε and IVC described above is a relationship based on the lower limit temperature of the combustion chamber 17 capable of realizing SPCCI combustion in layer 3 when the engine 1 is operated with a light load.

On the other hand, regardless of the layer 2 and the layer 3, if the temperature in the combustion chamber 17 is too high, CI combustion starts before the start of SI combustion, and SPCCI combustion cannot be performed.

Here, as described above, the combustion concept of SPCCI combustion is that when the spark plug 25 ignites the air-fuel mixture, the air-fuel mixture around the spark plug 25 starts SI combustion, and then the surrounding air-fuel mixture undergoes CI combustion. .. From the examination of experiments and the like conducted by the inventors of the present application, self-ignition of the air-fuel mixture occurs when the ambient temperature of the self-igniting air-fuel mixture exceeds a predetermined reference temperature Tth4, and the reference temperature Tth4 at this time is It was found that the temperature does not always match the average temperature of the entire combustion chamber 17. From this finding, it is considered that when the average temperature in the combustion chamber 17 at the compression top dead center reaches the reference temperature Tth4, CI combustion starts before SI combustion starts. In this case, SPCCI combustion is performed. I can't do it.

Therefore, the inventors of the present application have the valve closing timing IVC of the intake valve 21 and the intake valve 21 for each geometric compression ratio ε (ε1, ε2, ...), Similar to the matrix of reference numeral 1901 in FIG. And at the time of compression top dead center using the model of engine 1 while changing the values for each of the valve closing timing IVC and the overlap period O / L in the matrix with the overlap period O / L of the exhaust valve 22. The temperature of the combustion chamber 17 was estimated. The combination of IVC and O / L in which the temperature in the combustion chamber 17 at the compression top dead center exceeds the reference temperature Tth4 cannot realize SPCCI combustion, and the IVC and O / L having the reference temperature Tth4 or less cannot be realized. The combination of can realize SPCCI combustion.

FIG. 20 shows approximate equations (VII) and (VIII) calculated from the combination of ε and IVC in which the temperature in the combustion chamber 17 at the compression top dead center is equal to or less than the reference temperature Tth4. Approximate equations (VII) and (VIII) are, respectively.
Approximate formula (VII) IVC = -4.7481ε ² + 266.75ε-3671.2
And approximate formula (VIII) IVC = 4.7481ε ^2-266.75ε ＋ 3711.2
Is.

In FIG. 20, the combination of ε and IVC on the left side of the approximate equations (VII) and (VIII) can prevent CI combustion from starting before SI combustion, and SPCCI combustion is realized. ..

As can be seen from FIG. 20, also in layer 3, the relationship between ε and IVC is substantially symmetrical in the vertical direction with IVC = about 20 deg. A BDC as the center. Further, IVC = 75deg. ABDC is a retard limit of the valve closing timing of the intake valve 21 set in consideration of the amount of gas introduced into the combustion chamber 17 when the engine 1 is operated in layer 3, and IVC = Similarly, −35 deg. ABDC is the advance limit of the valve closing timing of the intake valve 21 set in consideration of the amount of gas introduced into the combustion chamber 17.

When the designer determines the valve closing timing IVC of the intake valve 21 when the engine 1 is operated in layer 3, ε surrounded by the approximate equations (V) (VI) (VII) (VIII) in FIG. -IVC must be defined within the range of establishment of IVC (the range shaded in FIG. 20).

Specifically, the designer sets the geometric compression ratio ε to 10 ≦ ε <20, and when the rotation speed of the engine 1 is 2000 rpm,
0.9949ε ² －41.736ε ＋ 401.16 ≦ IVC ≦ －0.9949ε ² ＋ 41.736ε－361.16… (11)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

The designer also sets the geometric compression ratio ε to 20 ≤ ε <25 when the engine speed is 2000 rpm.
-35 ≤ IVC ≤ 75 ... (12)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, the designer sets the geometric compression ratio ε to 25 ≦ ε ≦ 30, and when the rotation speed of the engine 1 is 2000 rpm,
－4.7481ε ² ＋266.75ε－3671.2 ≦ IVC ≦ 75… (13)
Or,
-35 ≤ IVC ≤ 4.7481ε ² -266.75ε + 3711.2 ... (14)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

By setting the valve closing timing IVC of the intake valve 21 based on the relational expressions (11) to (14), it is realized that the A / F burns the air-fuel mixture leaner than the stoichiometric air-fuel ratio by SPCCI combustion. The valve closing timing IVC is set for each operating state determined by the load and the rotation speed of the engine 1 in the layer 3.

As described above, the example shown by the solid line in FIG. 20 is the ε-IVC establishment range when the rotation speed of the engine 1 is 2000 rpm. When the number of revolutions of the engine 1 changes, the range of establishment of ε-IVC also changes. The point that the ε-IVC establishment range translates to the retard side as the rotation speed of the engine 1 increases is the same in FIG. Therefore, when the geometric compression ratio ε is set to 10 ≦ ε <20, the designer determines that the engine 1 has a rotation speed of 3000 rpm (see the broken line).
0.9949ε ² -41.736ε + 403.16 ≤ IVC ≤ -0.9949ε ² + 41.736ε-359.16… (11 ^-1 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

The designer also sets the geometric compression ratio ε to 20 ≤ ε <25 when the engine speed is 3000 rpm.
-33 ≤ IVC ≤ 77 ... (12 ^-1 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, when the designer sets the geometric compression ratio ε to 25 ≦ ε ≦ 30, when the rotation speed of the engine 1 is 3000 rpm,
－4.7481ε ² ＋266.75ε－3669.2 ≦ IVC ≦ 77… (13 ^-1 )
Or,
-33 ≤ IVC ≤ 4.7481ε ² -266.75ε + 3713.2 ... (14 ^-1 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

The designer also sets the geometric compression ratio ε to 10 ≤ ε <20, and when the engine 1 speed is 4000 rpm (see alternate long and short dash line),
0.9949ε ² -41.736ε + 409.16 ≤ IVC ≤ -0.9949ε ² + 41.736ε-353.16… (11 ^-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

The designer also sets the geometric compression ratio ε to 20 ≤ ε <25 when the engine speed is 4000 rpm.
-27 ≤ IVC ≤ 83 ... (12 ^-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, when the designer sets the geometric compression ratio ε to 25 ≦ ε ≦ 30, when the rotation speed of the engine 1 is 4000 rpm,
－4.7481ε ² ＋266.75ε－ ^3663.2 ≦ IVC ≦ 83… (13-2)
Or,
-27 ≤ IVC ≤ 4.7481ε ² -266.75ε + 3719.2 ... (14 ^-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Similar to the above, using the correction term C relating to the rotation speed NE (rpm) of the engine 1, the relational expression between ε and IVC in layer 3 can be expressed as follows.
If the geometric compression ratio ε is 10 ≤ ε <20,
0.9949ε ² －41.736ε ＋ 401.16 ＋ C ≦ IVC ≦ －0.9949ε ² ＋ 41.736ε－361.16 ＋ C… (11 ^-3 )
If the geometric compression ratio ε is 20 ≤ ε <25,
-35 + C ≤ IVC ≤ 75 + C ... (12 ^-3 )
If the geometric compression ratio ε is 25 ≤ ε ≤ 30
－4.7481ε ² ＋266.75ε－3671.2 ＋ C ≦ IVC ≦ 75 ＋ C… (13 ^-3 )
Or,
-35 ＋ C ≦ IVC ≦ 4.7481ε ² －266.75ε ＋ 3711.2 ＋ C… (14 ^-3 )
The designer determines the valve closing timing IVC based on the ε-IVC establishment range determined for each rotation speed of the engine 1. As a result, the designer can set the valve timing of the intake valve 21 in layer 3 as illustrated in FIG.

Further, FIG. 2002 in the lower part of FIG. 20 shows the relationship between ε and IVC when the fuel is a low octane number fuel.

In the engine 1 with a low octane number fuel, when the valve closing timing IVC is set, the designer sets the geometric compression ratio ε to 10 ≦ ε <18.7 when the rotation speed of the engine 1 is 2000 rpm.
0.9949ε ² －39.149ε ＋ 348.59 ≦ IVC ≦ －0.9949ε ² ＋ 39.149ε－308.59… (15)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, when the geometric compression ratio ε is set to 18.7 ≤ ε <23.7 in the engine 1 having a low octane number fuel, the designer determines that the rotation speed of the engine 1 is 2000 rpm.
-35 ≤ IVC ≤ 75 ... (16)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, in the engine 1 of the low octane number fuel, the designer sets the geometric compression ratio ε to 23.7 ≦ ε ≦ 30, and when the rotation speed of the engine 1 is 2000 rpm, the designer sets the geometric compression ratio ε to 23.7 ≦ ε ≦ 30.
－3.1298ε ² ＋172.48ε－2300 ≦ IVC ≦ 75… (17)
Or,
-35 ≤ IVC ≤ 3.1298ε ² -172.48ε + 2340 ... (18)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

The range of FIG. 20 with cross-hatching is the range in which the ε-IVC formation range of the high octane fuel and the ε-IVC formation range of the low octane fuel overlap. Similar to the above, when the IVC is defined within the range where the two establishment ranges overlap, the designer sets the control logic suitable for both the engine 1 using the high octane fuel and the engine 1 using the low octane fuel. be able to.

Although not shown, even in the engine 1 with a low octane number fuel, when the rotation speed of the engine 1 increases, the ε-IVC establishment range translates to the retard side. In the engine 1 with a low octane number fuel, the designer sets the geometric compression ratio ε to 10 ≦ ε <18.7, and when the rotation speed of the engine 1 is 3000 rpm, the designer sets it.
0.9949ε ² －39.149ε ＋ 350.59 ≦ IVC ≦ －0.9949ε ² ＋ 39.149ε－306.59… (15 ^-1 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, when the geometric compression ratio ε is set to 18.7 ≤ ε <23.7 in the engine 1 having a low octane number fuel, the designer determines that the rotation speed of the engine 1 is 3000 rpm.
-33 ≤ IVC ≤ 77 ... (16 ^-1 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, in the engine 1 of the low octane number fuel, the designer sets the geometric compression ratio ε to 23.7 ≦ ε ≦ 30, and when the rotation speed of the engine 1 is 3000 rpm,
－3.1298ε ² ＋172.48ε－2298 ≦ IVC ≦ 77… (17 ^-1 )
Or,
-33 ≤ IVC ≤ 3.1298ε ² -172.48ε + 2342 ... (18 ^-1 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

In the engine 1 with a low octane number fuel, the designer sets the geometric compression ratio ε to 10 ≦ ε <18.7, and when the rotation speed of the engine 1 is 4000 rpm, the designer sets it.
0.9949ε ² －39.149ε ＋356.59 ≦ IVC ≦ －0.9949ε ² ＋39.149ε－ ^300.59 … (15-2)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, when the geometric compression ratio ε is set to 18.7 ≤ ε <23.7 in the engine 1 having a low octane number fuel, the designer determines that the rotation speed of the engine 1 is 4000 rpm.
-27 ≤ IVC ≤ 83 ... ( ^16-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Further, in the engine 1 of the low octane number fuel, the designer sets the geometric compression ratio ε to 23.7 ≦ ε ≦ 30, and when the rotation speed of the engine 1 is 4000 rpm,
-3.1298ε ² + 172.48ε-2292 ≤ IVC ≤ 83 ... (17 ^-2 )
Or,
-27 ≤ IVC ≤ 3.1298 ε ² -172.48 ε + 2348 ... (18 ^-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Similarly to the above, when the correction term C relating to the rotation speed NE (rpm) of the engine 1 is used, the relational expression between ε and IVC in layer 3 can be expressed as follows in the engine 1 having a low octane number fuel. can.
If the geometric compression ratio ε is 10 ≤ ε <18.7,
0.9949ε ² －39.149ε ＋ 348.59 ＋ C ≦ IVC ≦ －0.9949ε ² ＋ 39.149ε－308.59 ＋ C… (15 ^-3 )
If the geometric compression ratio ε is 18.7 ≤ ε <23.7,
-35 + C ≤ IVC ≤ 75 + C ... (16 ^-3 )
If the geometric compression ratio ε is 23.7 ≤ ε ≤ 30
－3.1298ε ² ＋172.48ε－2300 ＋ C ≦ IVC ≦ 75 ＋ C… (17 ^-3 )
Or,
-35 + C ≤ IVC ≤ 3.1298ε ² -172.48ε + 2340 + C ... (18 ^-3 )
(Relationship between the geometric compression ratio in layer 2 and layer 3 and the valve closing timing of the intake valve)
FIG. 21 shows the relationship between the geometric compression ratio ε and the valve closing timing IVC of the intake valve 21, which enables SPCCI combustion in both layer 2 and layer 3. This relational expression is obtained from the ε-IVC establishment range of FIG. 17 and the ε-IVC establishment range of FIG.

When the ECU 10 selects the layer 3 according to the temperature of the engine 1 or the like, the low load operating region of the engine 1 is switched from the layer 2 to the layer 3. If the valve closing timing IVC of the intake valve 21 is set so that SPCCI combustion is possible in both layer 2 and layer 3, even when the map of the engine 1 is switched from layer 2 to layer 3, SPCCI is performed. It becomes possible to continuously carry out combustion.

FIG. 2101 above FIG. 21 shows the relationship between ε and IVC when the fuel is a high octane fuel. FIG. 2102 below shows the relationship between ε and IVC when the fuel is a low octane number fuel.

In the engine 1 with a high octane number fuel, the designer sets the geometric compression ratio ε to 10 ≦ ε <17, and when the rotation speed of the engine 1 is 2000 rpm, the designer sets it.
0.9949ε ² －41.736ε ＋ 401.16 ≦ IVC ≦ －0.9949ε ² ＋ 41.736ε－361.16… (19)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

In the engine 1 with a high octane number fuel, the designer sets the geometric compression ratio ε to 17 ≦ ε ≦ 30, and when the rotation speed of the engine 1 is 2000 rpm, the designer sets the geometric compression ratio ε to 17 ≦ ε ≦ 30.
－0.4288ε ² ＋31.518ε－379.88 ≦ IVC ≦ －0.9949ε ² ＋41.736ε－361.16… (20)
Or
0.9949ε ² －41.736ε ＋ 401.16 ≦ IVC ≦ 1.9163ε ² －89.935ε ＋ 974.94… (21)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

By setting the valve closing timing IVC of the intake valve 21 based on the relational expressions (19) to (21), the air-fuel mixture leaner than the stoichiometric air-fuel ratio can be burned by SPCCI. It is possible to SPCCI burn an air-fuel mixture having an A / F richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio and a leaner G / F than the stoichiometric air-fuel ratio.

The valve closing timing IVC is set for each operating state determined by the load and the rotation speed of the engine 1 in the layer 2 and the layer 3.

As shown by the broken line, when the geometric compression ratio ε is set to 10 ≦ ε <17 in the engine 1 having a high octane number fuel, when the rotation speed of the engine 1 is 3000 rpm, the designer sets the geometric compression ratio ε to 10 ≦ ε <17.
0.9949ε ² -41.736ε + 403.16 ≤ IVC ≤ -0.9949ε ² + 41.736ε-359.16… (19 ^-1 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

In the engine 1 with a high octane number fuel, the designer sets the geometric compression ratio ε to 17 ≦ ε ≦ 30, and when the rotation speed of the engine 1 is 3000 rpm, the designer sets the geometric compression ratio ε to 17 ≦ ε ≦ 30.
－0.4288ε ² ＋31.518ε－377.88 ≦ IVC ≦ －0.9949ε ² ＋41.736ε－359.16… (20 ^-1 )
Or
0.9949ε ² －41.736ε ＋ 403.16 ≦ IVC ≦ 1.9163ε ² －89.935ε ＋ 976.94… (21 ^-1 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Similarly, as shown by the alternate long and short dash line, the designer also sets the geometric compression ratio ε to 10 ≦ ε <17 in the engine 1 of the high octane number fuel, when the rotation speed of the engine 1 is 4000 rpm.
0.9949ε ² -41.736ε + 409.16 ≤ IVC ≤ -0.9949ε ² + 41.736ε- ^353.16 … (19-2)
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

In the engine 1 with a high octane number fuel, the designer sets the geometric compression ratio ε to 17 ≦ ε ≦ 30, and when the rotation speed of the engine 1 is 4000 rpm, the designer sets the geometric compression ratio ε to 17 ≦ ε ≦ 30.
－0.4288ε ² ＋31.518ε－371.88 ≦ IVC ≦ －0.9949ε ² ＋41.736ε－ ^353.16 … (20-2)
Or
0.9949ε ² －41.736ε ＋ 409.16 ≦ IVC ≦ 1.9163ε ² －89.935ε ＋ 982.94… (21 ^-2 )
The valve closing timing IVC (deg. ABDC) is set so as to satisfy.

Similar to the above, using the correction term C relating to the rotation speed NE (rpm) of the engine 1, the relational expression between ε and IVC in layer 2 and layer 3 can be expressed as follows.
If the geometric compression ratio ε is 10 ≤ ε <17,
0.9949ε ² －41.736ε ＋ 401.16 ＋ C ≦ IVC ≦ －0.9949ε ² ＋ 41.736ε－361.16 ＋ C… (19 ^-3 )
If the geometric compression ratio ε is 17 ≤ ε ≤ 30
－0.4288ε ² ＋31.518ε－379.88 ＋ C ≦ IVC ≦ －0.9949ε ² ＋41.736ε－361.16 ＋ C… (20 ^-3 )
Or
0.9949ε ² －41.736ε ＋ 401.16 ＋ C ≦ IVC ≦ 1.9163ε ² －89.935ε ＋ 974.94 ＋ C… (21 ^-3 )
Here, if the geometric compression ratio ε is set to less than 17, the designer can determine the IVC based on the relational expression (19 ^-3 ). Since the selection range of IVC is wide, the degree of freedom in design is high.

Further, as shown in FIG. 212 below, when the designer sets the geometric compression ratio ε to 10 ≦ ε <15.7 in the engine 1 having a low octane number fuel, the rotation speed of the engine 1 is 2000 rpm. sometimes,
0.9949ε ² －39.149ε ＋ 348.59 ≦ IVC ≦ －0.9949ε ² ＋ 39.149ε－308.59… (22)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Further, when the designer sets the geometric compression ratio ε to 15.7 ≦ ε ≦ 30 in the engine 1 having a low octane number fuel, when the rotation speed of the engine 1 is 2000 rpm, the designer sets the geometric compression ratio ε to 15.7 ≦ ε ≦ 30.
-0.5603ε ² ＋34.859ε－377.22 ≦ IVC ≦ －0.9949ε ² ＋39.149ε－308.59… (23)
Or,
0.9949ε ² －39.149ε ＋ 348.59 ≦ IVC ≦ 1.9211ε ² -85.076ε ＋ 862.01… (24)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Although not shown, the designer sets the geometric compression ratio ε to 10 ≦ ε <15.7 in the engine 1 with a low octane number fuel, and when the rotation speed of the engine 1 is 3000 rpm,
0.9949ε ² －39.149ε ＋ 350.59 ≦ IVC ≦ －0.9949ε ² ＋ 39.149ε－306.59… (22 ^-1 )
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Further, when the designer sets the geometric compression ratio ε to 15.7 ≦ ε ≦ 30 in the engine 1 having a low octane number fuel, when the rotation speed of the engine 1 is 3000 rpm, the designer sets the geometric compression ratio ε to 15.7 ≦ ε ≦ 30.
-0.5603ε ² ＋34.859ε－375.22 ≦ IVC ≦ －0.9949ε ² ＋39.149ε－306.59… (23 ^-1 )
Or,
0.9949ε ² －39.149ε ＋ 350.59 ≦ IVC ≦ 1.9211ε ² -85.076ε ＋ 864.01… (24 ^-1 )
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Similarly, although not shown, the designer sets the geometric compression ratio ε to 10 ≦ ε <15.7 in the engine 1 with a low octane number fuel, and when the rotation speed of the engine 1 is 4000 rpm,
0.9949ε ² －39.149ε ＋356.59 ≦ IVC ≦ －0.9949ε ² ＋39.149ε－300.59… (22 ^-2 )
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Further, when the designer sets the geometric compression ratio ε to 15.7 ≦ ε ≦ 30 in the engine 1 having a low octane number fuel, when the rotation speed of the engine 1 is 4000 rpm, the designer sets the geometric compression ratio ε to 15.7 ≦ ε ≦ 30.
-0.5603ε ² ＋34.859ε－369.22 ≦ IVC ≦ －0.9949ε ² ＋39.149ε－300.59… (23 ^-2 )
Or,
0.9949ε 2-39.149ε ＋356.59 ≦ IVC ≦ ^1.9211ε 2-85.076ε ＋ ^870.01 … ( ^24-2 )
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Similarly to the above, when the correction term C relating to the rotation speed NE (rpm) of the engine 1 is used, in the engine 1 having a low octane number, the relational expression between ε and IVC in layer 2 and layer 3 is as follows. Can be represented.
If the geometric compression ratio ε is 10 ≦ ε <15.7,
0.9949ε ² －39.149ε ＋ 348.59 ＋ C ≦ IVC ≦ －0.9949ε ² ＋ 39.149ε－308.59 ＋ C… (22 ^-3 )
If the geometric compression ratio ε is 15.7 ≤ ε ≤ 30
-0.5603ε ² ＋34.859ε－377.22 ＋ C ≦ IVC ≦ －0.9949ε ² ＋39.149ε－308.59 ＋ C… (23 ^-3 )
Or,
0.9949ε ² －39.149ε ＋ 348.59 ＋ C ≦ IVC ≦ 1.9211ε ² -85.076ε ＋ 862.01 ＋ C… (24 ^-3 )
Although not shown, the designer may determine the IVC in the range where the ε-IVC formation range in FIG. 21 above and the ε-IVC formation range in FIG. 2102 below overlap. Similar to the above, when the IVC is defined within the range where the two establishment ranges overlap, the designer sets the control logic suitable for both the engine 1 using the high octane fuel and the engine 1 using the low octane fuel. be able to.

(Procedure of control logic design method)
Next, the procedure of the method of designing the control logic of the engine 1 which performs SPCCI combustion will be described with reference to the flowchart shown in FIG. The designer can use a computer to perform each step. The computer stores the information regarding the ε-IVC establishment range exemplified in FIGS. 17, 20, and 21.

First, in step S221 after the start, the designer sets the geometric compression ratio ε. The designer may input the value of the set geometric compression ratio ε into the computer.

In the following step S222, the designer sets the valve opening angle of the intake valve 21 and the valve opening angle of the exhaust valve 22, respectively. This corresponds to defining the cam shape of the intake valve 21 and the cam shape of the exhaust valve 22. The designer may input the values of the set valve opening angle of the intake valve 21 and the valve opening angle of the exhaust valve 22 into the computer. In steps S221 and S222, the hardware configuration of the engine 1 can be set.

In step S223, the designer sets the operating state consisting of the load and the rotation speed of the engine 1, and in the following step S224, the designer sets the ε-IVC establishment range (FIGS. 17, 20 and 21) stored in the computer. ), Select IVC.

Then, in step S225, the computer determines whether or not SPCCI combustion is feasible based on the IVC set in step S224. If the determination in step S225 is YES, the process proceeds to step S226, and the designer determines the control logic of the engine 1 so as to execute SPCCI combustion in the operating state set in step S223. On the other hand, if the determination in step S225 is NO, the process proceeds to step S227, and the designer determines the control logic of the engine 1 so as to execute SI combustion in the operating state set in step S223. In step S227, the designer may set the valve closing timing IVC of the intake valve 21 again in consideration of performing SI combustion.

As described above, the method of designing the control logic of the compression ignition type engine disclosed here defines the relationship between the geometric compression ratio ε of the engine and the valve closing timing IVC of the intake valve 21. The designer can determine the valve closing timing IVC of the intake valve 21 within a range that satisfies the relationship. The designer can design the control logic of the engine 1 with less man-hours than in the past.

(Other embodiments)
The technique disclosed herein is not limited to being applied to the engine 1 having the above-described configuration. As the configuration of the engine 1, various configurations can be adopted.

For example, the engine 1 may be provided with a turbocharger instead of the mechanical supercharger 44.

1 Engine 10 ECU (control unit)
17 Combustion chamber 23 Intake electric S-VT (variable valve mechanism)
25 Spark plug (ignition part)
44 Supercharger 55 EGR system 6 Injector (fuel injection part)
SW1 Airflow sensor (measurement unit)
SW2 1st intake air temperature sensor (measurement unit)
SW3 1st pressure sensor (measurement unit)
SW4 2nd intake air temperature sensor (measurement unit)
SW5 2nd pressure sensor (measurement unit)
SW6 Shiatsu sensor (measurement unit)
SW7 Exhaust temperature sensor (measurement unit)
SW8 Linear O ₂ sensor (measurement unit)
SW9 Lambda O ₂ sensor (measurement unit)
SW10 water temperature sensor (measurement unit)
SW11 Crank angle sensor (measurement unit)
SW12 Accelerator opening sensor (measurement unit)
SW13 Intake cam angle sensor (measurement unit)
SW14 Exhaust cam angle sensor (measurement unit)
SW15 EGR differential pressure sensor (measurement unit)
SW16 Fuel pressure sensor (measurement unit)
SW17 3rd intake air temperature sensor (measurement unit)

Claims (13)

It is a method of designing the control logic of the compression ignition type engine installed in the automobile .
The engine
A fuel injection unit that injects fuel to be supplied into the combustion chamber partitioned by a cylinder and a piston ,
A variable valve mechanism that changes the valve timing of the intake valve that opens and closes the combustion chamber ,
An ignition unit that ignites the air-fuel mixture in the combustion chamber,
A measuring unit that measures parameters related to the operating state of the engine,
A control unit that receives the measurement results of the measurement unit, performs calculations according to the control logic corresponding to the operating state of the engine, and outputs a signal to the fuel injection unit, the variable valve mechanism, and the ignition unit. , Equipped with
In the control unit, between a predetermined light load and an upper limit load, the G / F, which is the weight ratio of the total gas of the air-fuel mixture in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. A signal is output to each of the fuel injection unit and the variable valve mechanism so that the A / F, which is the weight ratio of the air-fuel mixture to the fuel, becomes richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio. After the ignition unit ignites the air-fuel mixture in the combustion chamber, a signal is output to the ignition unit so that the unburned air-fuel mixture burns by self-ignition.
The method of designing the control logic is
The step of determining the geometric compression ratio ε of the engine and
A step of determining a control logic that defines the valve closing timing IVC of the intake valve is provided.
When the control logic is set, if the fuel is a high octane fuel and the geometric compression ratio ε is 10 ≦ ε <17.
0.4234ε ^2-22.926ε ＋ 207.84 ＋ C ≦ IVC ≦ －0.4234ε ² ＋ 22.926ε－167.84 ＋ C
However, C is a correction term related to the engine speed NE (rpm).
C = 3.3 × 10 ^-10 NE ³ －1.0 × 10 ^-6 NE ² + 7.0 × 10 ^-4 NE
A method of designing a control logic of a compression ignition type engine that determines the valve closing timing IVC (deg. ABDC) so as to satisfy the above. It is a method of designing the control logic of the compression ignition type engine installed in the automobile .
The engine
A fuel injection unit that injects fuel to be supplied into the combustion chamber partitioned by a cylinder and a piston ,
A variable valve mechanism that changes the valve timing of the intake valve that opens and closes the combustion chamber ,
An ignition unit that ignites the air-fuel mixture in the combustion chamber,
A measuring unit that measures parameters related to the operating state of the engine,
A control unit that receives the measurement results of the measurement unit, performs calculations according to the control logic corresponding to the operating state of the engine, and outputs a signal to the fuel injection unit, the variable valve mechanism, and the ignition unit. , Equipped with
In the control unit, between a predetermined light load and an upper limit load, the G / F, which is the weight ratio of the total gas of the air-fuel mixture in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. A signal is output to each of the fuel injection unit and the variable valve mechanism so that the A / F, which is the weight ratio of the air-fuel mixture to the fuel, becomes richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio. After the ignition unit ignites the air-fuel mixture in the combustion chamber, a signal is output to the ignition unit so that the unburned air-fuel mixture burns by self-ignition.
The method of designing the control logic is
The step of determining the geometric compression ratio ε of the engine and
A step of determining a control logic that defines the valve closing timing IVC of the intake valve is provided.
When the control logic is set, if the fuel is a high octane fuel and the geometric compression ratio ε is 17 ≦ ε <20.
-0.4288ε ² + 31.518ε-379.88 + C ≤ IVC ≤ -0.4234ε ² + 22.926ε-167.84 + C
Or,
0.4234ε 2-22.926ε ＋ 207.84 ＋ C ≦ IVC ≦ 1.9163ε ^{2 －89.935ε} ＋ 974.94 ＋ C
However, C is a correction term related to the engine speed NE (rpm).
C = 3.3 × 10 ^-10 NE ³ －1.0 × 10 ^-6 NE ² + 7.0 × 10 ^-4 NE
A method of designing a control logic of a compression ignition type engine that determines the valve closing timing IVC (deg. ABDC) so as to satisfy the above. It is a method of designing the control logic of the compression ignition type engine installed in the automobile .
The engine
A fuel injection unit that injects fuel to be supplied into the combustion chamber partitioned by a cylinder and a piston ,
A variable valve mechanism that changes the valve timing of the intake valve that opens and closes the combustion chamber ,
An ignition unit that ignites the air-fuel mixture in the combustion chamber,
A measuring unit that measures parameters related to the operating state of the engine,
A control unit that receives the measurement results of the measurement unit, performs calculations according to the control logic corresponding to the operating state of the engine, and outputs a signal to the fuel injection unit, the variable valve mechanism, and the ignition unit. , Equipped with
In the control unit, between a predetermined light load and an upper limit load, the G / F, which is the weight ratio of the total gas of the air-fuel mixture in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. A signal is output to each of the fuel injection unit and the variable valve mechanism so that the A / F, which is the weight ratio of the air-fuel mixture to the fuel, becomes richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio. After the ignition unit ignites the air-fuel mixture in the combustion chamber, a signal is output to the ignition unit so that the unburned air-fuel mixture burns by self-ignition.
The method of designing the control logic is
The step of determining the geometric compression ratio ε of the engine and
A step of determining a control logic that defines the valve closing timing IVC of the intake valve is provided.
When the control logic is set, if the fuel is a high octane fuel and the geometric compression ratio ε is 20 ≦ ε ≦ 30.
-0.4288ε ² + 31.518ε-379.88 + C ≤ IVC ≤ 120
Or,
-80 ≤ IVC ≤ 1.9163ε ² -89.935ε + 974.94 + C
However, C is a correction term related to the engine speed NE (rpm).
C = 3.3 × 10 ^-10 NE ³ －1.0 × 10 ^-6 NE ² + 7.0 × 10 ^-4 NE
A method of designing a control logic of a compression ignition type engine that determines the valve closing timing IVC (deg. ABDC) so as to satisfy the above. It is a method of designing the control logic of the compression ignition type engine installed in the automobile .
The engine
A fuel injection unit that injects fuel to be supplied into the combustion chamber partitioned by a cylinder and a piston ,
A variable valve mechanism that changes the valve timing of the intake valve that opens and closes the combustion chamber ,
An ignition unit that ignites the air-fuel mixture in the combustion chamber,
A measuring unit that measures parameters related to the operating state of the engine,
A control unit that receives the measurement results of the measurement unit, performs calculations according to the control logic corresponding to the operating state of the engine, and outputs a signal to the fuel injection unit, the variable valve mechanism, and the ignition unit. , Equipped with
In the control unit, between a predetermined light load and an upper limit load, the G / F, which is the weight ratio of the total gas of the air-fuel mixture in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. A signal is output to each of the fuel injection unit and the variable valve mechanism so that the A / F, which is the weight ratio of the air-fuel mixture to the fuel, becomes richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio. After the ignition unit ignites the air-fuel mixture in the combustion chamber, a signal is output to the ignition unit so that the unburned air-fuel mixture burns by self-ignition.
The method of designing the control logic is
The step of determining the geometric compression ratio ε of the engine and
A step of determining a control logic that defines the valve closing timing IVC of the intake valve is provided.
When the control logic is set, if the fuel is a low octane fuel and the geometric compression ratio ε is 10 ≦ ε <15.7.
0.4234ε ² －21.826ε ＋ 178.75 ＋ C ≦ IVC ≦ －0.4234ε ² ＋ 21.826ε－138.75 ＋ C
However, C is a correction term related to the engine speed NE (rpm).
C = 3.3 × 10 ^-10 NE ³ －1.0 × 10 ^-6 NE ² + 7.0 × 10 ^-4 NE
A method of designing a control logic of a compression ignition type engine that determines the valve closing timing IVC (deg. ABDC) so as to satisfy the above. It is a method of designing the control logic of the compression ignition type engine installed in the automobile .
The engine
A fuel injection unit that injects fuel to be supplied into the combustion chamber partitioned by a cylinder and a piston ,
A variable valve mechanism that changes the valve timing of the intake valve that opens and closes the combustion chamber ,
An ignition unit that ignites the air-fuel mixture in the combustion chamber,
A measuring unit that measures parameters related to the operating state of the engine,
A control unit that receives the measurement results of the measurement unit, performs calculations according to the control logic corresponding to the operating state of the engine, and outputs a signal to the fuel injection unit, the variable valve mechanism, and the ignition unit. , Equipped with
In the control unit, between a predetermined light load and an upper limit load, the G / F, which is the weight ratio of the total gas of the air-fuel mixture in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. A signal is output to each of the fuel injection unit and the variable valve mechanism so that the A / F, which is the weight ratio of the air-fuel mixture to the fuel, becomes richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio. After the ignition unit ignites the air-fuel mixture in the combustion chamber, a signal is output to the ignition unit so that the unburned air-fuel mixture burns by self-ignition.
The method of designing the control logic is
The step of determining the geometric compression ratio ε of the engine and
A step of determining a control logic that defines the valve closing timing IVC of the intake valve is provided.
When the control logic is set, if the fuel is a low octane fuel and the geometric compression ratio ε is 15.7 ≦ ε <18.7.
-0.5603ε ² + 34.859ε-377.22 + C ≤ IVC ≤ -0.4234ε ² + 21.826ε-138.75 + C
Or,
0.4234ε ² －21.826ε ＋ 178.75 ＋ C ≦ IVC ≦ 1.9211ε ² -85.076ε ＋ 862.01 ＋ C
However, C is a correction term related to the engine speed NE (rpm).
C = 3.3 × 10 ^-10 NE ³ －1.0 × 10 ^-6 NE ² + 7.0 × 10 ^-4 NE
A method of designing a control logic of a compression ignition type engine that determines the valve closing timing IVC (deg. ABDC) so as to satisfy the above. It is a method of designing the control logic of the compression ignition type engine installed in the automobile .
The engine
A fuel injection unit that injects fuel to be supplied into the combustion chamber partitioned by a cylinder and a piston ,
A variable valve mechanism that changes the valve timing of the intake valve that opens and closes the combustion chamber ,
An ignition unit that ignites the air-fuel mixture in the combustion chamber,
A measuring unit that measures parameters related to the operating state of the engine,
A control unit that receives the measurement results of the measurement unit, performs calculations according to the control logic corresponding to the operating state of the engine, and outputs a signal to the fuel injection unit, the variable valve mechanism, and the ignition unit. , Equipped with
In the control unit, between a predetermined light load and an upper limit load, the G / F, which is the weight ratio of the total gas of the air-fuel mixture in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. A signal is output to each of the fuel injection unit and the variable valve mechanism so that the A / F, which is the weight ratio of the air-fuel mixture to the fuel, becomes richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio. After the ignition unit ignites the air-fuel mixture in the combustion chamber, a signal is output to the ignition unit so that the unburned air-fuel mixture burns by self-ignition.
The method of designing the control logic is
The step of determining the geometric compression ratio ε of the engine and
A step of determining a control logic that defines the valve closing timing IVC of the intake valve is provided.
When the control logic is set, if the fuel is a low octane fuel and the geometric compression ratio ε is 18.7 ≦ ε ≦ 30.
-0.5603ε ² ＋34.859ε－377.22 ＋ C ≦ IVC ≦ 120
Or,
-80 ≤ IVC ≤ 1.9211ε ^2-85.076ε + 862.01 + C
However, C is a correction term related to the engine speed NE (rpm).
C = 3.3 × 10 ^-10 NE ³ －1.0 × 10 ^-6 NE ² + 7.0 × 10 ^-4 NE
A method of designing a control logic of a compression ignition type engine that determines the valve closing timing IVC (deg. ABDC) so as to satisfy the above. In the method for designing the control logic of the compression ignition type engine according to any one of claims 1 to 6.
The valve closing timing IVC of the intake valve changes when the operating state of the engine changes.
A method of designing a control logic of a compression ignition type engine that determines the valve closing timing IVC (deg. ABDC) so as to satisfy the above relational expression for each operating state. In the method for designing the control logic of the compression ignition type engine according to any one of claims 1 to 7.
The upper limit load is a high load equal to or higher than a predetermined load.
A method of designing a control logic of a compression ignition type engine in which the engine operates in a high load operating state of a predetermined load or more according to the control logic. In the method of designing the control logic of the compression ignition type engine according to claim 8.
The upper limit load is the maximum load.
The engine is a method of designing a control logic of a compression ignition type engine that operates in a maximum load operating state according to the control logic. In the method of designing the control logic of the compression ignition type engine according to claim 8 or 9.
The engine is equipped with a supercharger.
The control unit controls a compression ignition type engine that outputs a signal to the supercharger so that the supercharger supercharges when the engine is in a high load operation state of a predetermined load or more according to the control logic. How to design the logic. In the method for designing the control logic of the compression ignition type engine according to any one of claims 1 to 10.
The control unit is a method of designing a control logic of a compression ignition type engine that makes the G / F of the air-fuel mixture 18 or more and 50 or less according to the control logic. In the method for designing the control logic of the compression ignition type engine according to any one of claims 1 to 11.
The engine comprises an EGR system that introduces exhaust gas into the combustion chamber.
The control unit determines the calorific value ratio as an index related to the ratio of the calorific value generated when the air-fuel mixture burns due to flame propagation to the total heat generated when the air-fuel mixture in the combustion chamber burns. A method of designing a control logic of a compression ignition type engine that outputs a signal to the EGR system and the ignition unit so as to have a target heat quantity ratio determined according to the operating state of the engine. In the method of designing the control logic of the compression ignition type engine according to claim 12.
The control unit is a method of designing a control logic of a compression ignition type engine that outputs a signal to the EGR system and the ignition unit so that the calorific value ratio becomes higher when the load of the engine is high than when the load is low.

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