JP2019203431A - Method of designing control logic of compression ignition type engine - Google Patents

Abstract

To provide a method of designing a control logic of a compression ignition type engine.SOLUTION: An engine 1 has a first mode in which an A/F is made lean, and after an ignition part ignites an air-fuel mixture in a combustion chamber, an unburned air-fuel mixture burns with self ignition, and a second mode in which a G/F is made lean and the A/F is made equal to or richer than a theoretical air-fuel ratio. A method of designing a control logic includes the steps of defining a geometric compression ratio ε, and determining a control logic defining a valve closing timing IVC of an intake valve. When the geometric compression ratio ε is 10≤ε<17, the IVC(deg. aBDC) is defined so as to satisfy 0.9949ε-41.736ε+401.16+C≤IVC≤-0.9949ε+41.736ε-361.16+C, in which C relates to a revolution speed NE(rpm) of the engine, and C=3.3×10NE-1.0×10NE+7.0×10NE.SELECTED DRAWING: Figure 21

Description

  The technology disclosed herein relates to a method for designing control logic for a compression ignition engine.

  It is known that the combustion by the compression self-ignition in which the air-fuel mixture burns at a time without passing through the flame propagation has the minimum combustion period, and thus maximizes the fuel efficiency. However, combustion by compression self-ignition needs to solve various problems in automobile engines. For example, in an automobile application, since driving conditions and environmental conditions change greatly, it is a big problem to stably perform compression self-ignition. In automobile engines, combustion by compression self-ignition has not yet been put into practical use. 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 hardly occurs. By performing ignition 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 that assists compression self-ignition by ignition of a spark plug, the applicant of the present application is an SPCCI (SPark Controlled Compression Ignition) that combines SI (Spark Ignition) combustion and CI (Compression Ignition) combustion. ) Propose combustion. SI combustion is combustion with flame propagation that starts by forcibly igniting the air-fuel mixture in the combustion chamber. CI combustion is combustion that starts when the air-fuel mixture in the combustion chamber undergoes compression self-ignition. 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 due to the heat generated by SI combustion and the pressure increase due to flame propagation. Is a form that burns by compression ignition. SPCCI combustion is a form of “combustion by compression ignition” because it includes CI combustion.

  CI combustion occurs when the in-cylinder temperature reaches an 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 the SPCCI combustion is a result of two pressure increases: a pressure increase due to the compression work of the piston in the compression stroke, and a pressure increase resulting from the heat generation of the 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 increase due to the compression work of the piston is small. In this case, the in-cylinder temperature cannot be increased to the ignition temperature unless the flame propagation in the SPCCI combustion progresses and the pressure increase caused by the heat generated by the SI combustion does not increase considerably. As a result, the amount of air-fuel mixture that undergoes compression self-ignition is small, and a large amount of air-fuel mixture is combusted by flame propagation, so the combustion period is long and fuel efficiency is reduced. That is, the effective compression ratio needs to be maintained at a certain value or more in order to stably cause CI combustion in SPCCI combustion and maximize fuel efficiency.

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

  In order for a designer to put an engine that performs SPCCI combustion into practical use, the minimum effective compression ratio that causes stable CI combustion is determined for the operating state of the engine, and the effective compression ratio is as long as combustion noise is allowed. However, it is necessary to determine an effective compression ratio so that 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 combustion noise within an allowable range.

  However, since SPCCI combustion is a new combustion method, no suitable effective compression ratio range has ever been found.

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

  The higher the geometric compression ratio, the higher the maximum in-cylinder pressure. Therefore, it is necessary to increase the strength of the engine component, which leads to an increase in weight and 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 varies depending on the heat balance based on the combustion mode and stroke volume even if the geometric compression ratio is the same, the optimum geometric compression ratio for the new combustion method SPCCI combustion However, it is unknown.

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

  Specifically, the technology disclosed herein relates to a method for designing control logic of a compression ignition type engine.

  The engine includes a fuel injection unit that injects fuel supplied into the combustion chamber, a variable valve mechanism that changes a valve timing of an intake valve, an ignition unit that ignites an air-fuel mixture in the combustion chamber, and an operating state of the engine A measurement unit that measures a parameter related to the above, a measurement result of the measurement unit, a calculation according to a control logic corresponding to the operating state of the engine, and the fuel injection unit, the variable valve mechanism, and A control unit that outputs a signal to the ignition unit.

  In the engine, the control unit has the fuel injection unit and the variable valve mechanism so that an A / F, which is a weight ratio between the air in the combustion chamber and the fuel, becomes leaner than the stoichiometric air-fuel ratio. A first mode for outputting a signal to each, and outputting a signal to the ignition unit so that the unburned mixture burns by self-ignition after the ignition unit ignites the mixture in the combustion chamber; and the control unit However, G / F, which is the weight ratio of the total gas of the air-fuel mixture in the combustion chamber and the fuel, is leaner than the stoichiometric air-fuel ratio, and the A / F is richer than the stoichiometric air-fuel ratio or stoichiometric air-fuel ratio. A signal is output to each of the fuel injection unit and the variable valve mechanism, and a signal is sent to the ignition unit so that the unburned mixture is burned by self-ignition after the ignition unit ignites the mixture in the combustion chamber. The second mode to output Having.

The method of designing the control logic includes the steps of determining a geometric compression ratio ε of the engine and determining control logic that determines a closing timing IVC of the intake valve, and setting the control logic When 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 (a)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.
Here, C is a correction term relating to the engine speed NE (rpm), and is C = 3.3 × 10 ⁻¹⁰ NE ³ −1.0 × 10 ⁻⁶ NE ³ + 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 mixture is combusted by self-ignition to complete the combustion. That is, this engine performs SPCCI combustion.

  If the engine is in the first mode, the A / F of the air-fuel mixture is made leaner than the stoichiometric air-fuel ratio. By making the A / F of the air-fuel mixture lean, the fuel efficiency of the engine is improved. In the second mode, the engine also makes G / F leaner than the stoichiometric air-fuel ratio and makes A / F richer than the stoichiometric air-fuel ratio or stoichiometric air-fuel ratio. By making G / F lean, fuel efficiency is improved. Moreover, by introducing the EGR gas into the combustion chamber, the combustion stability of SPCCI combustion is increased.

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

  So far, there has not been an index for determining the closing timing IVC of the intake valve that can be used when designing the control logic of an engine that performs SPCCI combustion. The designer had to determine the IVC corresponding to the operating state of the engine by repeatedly performing experiments and the like under various conditions while changing the 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 closing timing IVC of the intake valve in order to achieve appropriate SPCCI combustion. If the designer determines the valve closing timing IVC of the intake valve within a range that satisfies the relationship, the engine that performs SPCCI combustion can be put into practical use. The designer can put an engine that performs SPCCI combustion into practical use with very few man-hours.

When setting the control logic, 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 (b)
Or
0.9949ε ² −41.736ε + 401.16 + 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.

  Thereby, the engine that performs SPCCI combustion has the maximum fuel efficiency in each of the first mode and the second mode. Further, the designer can put an engine that performs SPCCI combustion into practical use with a small number of man-hours.

When setting the control logic, if the fuel is a low octane fuel and 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 (d)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

  This maximizes fuel efficiency even for engines that use low octane fuel. Further, the designer can put an engine that performs SPCCI combustion into practical use with a small number of man-hours.

  When the designer determines the valve closing timing IVC of the intake valve so as to satisfy both the expressions (a) and (d) when the geometric compression ratio ε is 10 ≦ ε <15.7, Control logic can be set that is compatible with both high octane fuel engines and low octane fuel engines. Even if the octane number of the fuel is different for each destination, the designer can design the engine control logic in a lump, and the design man-hour is reduced.

When setting the control logic, if the fuel is a low octane fuel and 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 (e)
Or
0.9949ε ² −39.149ε + 348.59 + C ≦ IVC ≦ 1.9211ε ² −85.076ε + 862.01 + C (f)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

  The closing timing IVC of the intake valve changes when the engine operating state changes, and the closing timing IVC of the intake valve is set so as to satisfy any one of the relational expressions (a) to (f). The valve timing IVC (deg. ABDC) may be determined.

  By doing so, the engine can stably perform SPCCI combustion while suppressing combustion noise in each of the first mode and the second mode in various operating states.

  The engine may be operated in a low load operation state of a predetermined load or less according to the control logic.

  In general CI combustion, the ignitability is reduced when the engine load is low. On the other hand, SPCCI combustion is SI combustion at the start of combustion, and CI combustion starts using heat generated by SI combustion. SPCCI combustion does not reduce ignitability even when the engine load is low.

  The engine may be operated in a minimum load operation state according to the control logic. That is, the engine may perform SPCCI combustion in the minimum load operation state.

  The controller is in a first mode when the wall temperature of the combustion chamber is equal to or higher than a first predetermined temperature and the temperature of the intake air is equal to or higher than a second predetermined temperature, and the wall temperature of the combustion chamber is set to the first predetermined temperature. When the temperature of the intake air is lower than the second predetermined temperature, the second mode may be set.

  In SPCCI combustion, part of the air-fuel mixture in the combustion chamber undergoes SI combustion, and the rest undergoes CI combustion. SPCCI combustion is affected by both the wall temperature of the combustion chamber and the temperature of the intake air. When the wall temperature of the combustion chamber and the temperature of the intake air are high, in the first mode, the air-fuel mixture whose A / F is leaner than the stoichiometric air-fuel ratio is stably subjected to SPCCI combustion while suppressing combustion noise. Can do. When the wall temperature of the combustion chamber and the temperature of the intake air are low, the temperature of the combustion chamber is increased by introducing high-temperature EGR gas into the combustion chamber in the second mode. As a result, the A / F is richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio, and the air-fuel mixture whose G / F is leaner than the stoichiometric air-fuel ratio can be stably subjected to SPCCI combustion while suppressing combustion noise. it can.

  The engine is provided with an EGR system that introduces exhaust gas into the combustion chamber, and the control unit is configured to mix the mixture by flame propagation with respect to the total amount of heat generated when the mixture in the combustion chamber burns. A signal is output to the EGR system and the ignition unit so that the heat quantity ratio as an index related to the ratio of the heat quantity generated when the engine burns becomes a target heat quantity ratio determined in accordance with the operating state of the engine. You may do it.

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

  When the heat quantity ratio is increased in SPCCI combustion, the SI combustion ratio is increased, which is advantageous for suppressing combustion noise. Lowering the heat quantity ratio in SPCCI combustion increases the CI combustion ratio, which is advantageous for improving fuel efficiency. The heat quantity ratio is changed 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 advanced, so the heat quantity ratio is low. Further, when the ignition timing is advanced, the SI combustion start timing is advanced, so that the heat quantity ratio is increased. The control unit outputs a signal to the EGR system and the ignition unit so that the heat amount ratio becomes a target heat amount ratio determined in accordance with the operating state of the engine, thereby reducing both combustion noise and improving fuel efficiency. .

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

  As the engine load increases, the amount of fuel supplied into the combustion chamber increases and the temperature in the combustion chamber increases. When the engine load is high, combustion noise can be suppressed by increasing the heat quantity ratio of SPCCI combustion.

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

FIG. 1 is a diagram illustrating a configuration of an engine. FIG. 2 is a diagram illustrating the configuration of the combustion chamber. The upper diagram is a plan view equivalent view of the combustion chamber, and the lower diagram is a sectional view taken along the 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 apparatus. FIG. 5 is a diagram illustrating a waveform of SPCCI combustion. FIG. 6 is a diagram illustrating a map of the engine. The upper diagram is a map during warm, the middle diagram is a map during semi-warm-up, and the lower diagram is a map during cold. FIG. 7 is a diagram illustrating details of a map during warm weather. FIG. 8 is a diagram illustrating fuel injection timing and ignition timing and combustion waveforms in each operation region of the map of FIG. FIG. 9 is a diagram for explaining a layer structure of an engine map. FIG. 10 is a flowchart illustrating a control process related to map layer selection. The upper diagram of FIG. 11 is a diagram illustrating the relationship between the engine load and intake valve opening timing in layer 2, and the lower diagram is a diagram illustrating the relationship between the engine speed and the intake valve opening timing in layer 2. The upper diagram of FIG. 12 shows the relationship between the engine load in layer 3 and the opening timing of the intake valve, the middle diagram shows the relationship between the engine load in layer 3 and the closing timing of the exhaust valve, and the lower diagram shows the engine in layer 3 It is a figure which illustrates the relationship between load and the overlap period of an intake valve and an exhaust valve. FIG. 13 is a flowchart illustrating an example of an engine operation control process executed by the ECU. FIG. 14 is a diagram for explaining the relationship between the engine load level and the target SI rate level. FIG. 15 is a diagram illustrating the establishment range of SPCCI combustion with respect to the EGR rate in the layer 2. FIG. 16 is an example of a matrix image used to determine the relationship between the geometric compression ratio at which SPCCI combustion is possible in layer 2 and the closing timing of the intake valve. The upper diagram of FIG. 17 shows the relationship between the geometric compression ratio at which SPCCI combustion is possible in Layer 2 and the closing timing of the intake valve when using a high octane fuel, and the lower diagram is when using a low octane fuel. This is the relationship between the geometric compression ratio at which SPCCI combustion is possible in layer 2 and the closing timing of the intake valve. FIG. 18 is a diagram illustrating 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 to determine the relationship between the geometric compression ratio at which SPCCI combustion is possible in layer 3 and the closing timing of the intake valve. The upper diagram of FIG. 20 shows the relationship between the geometric compression ratio at which SSPCCI combustion is possible in the layer 3 and the closing timing of the intake valve when the high octane fuel is used, and the lower diagram is when the low octane fuel is used. This is the relationship between the geometric compression ratio at which SPCCI combustion is possible in layer 3 and the closing timing of the intake valve. The upper diagram of FIG. 21 shows the relationship between the geometric compression ratio at which SPCCI combustion is possible in layer 2 and layer 3 and the closing timing of the intake valve when using a high octane fuel, and the lower diagram shows the low octane fuel Is the relationship between the geometric compression ratio at which SPCCI combustion is possible in layer 2 and layer 3 and the closing timing of the intake valve. FIG. 22 is a flowchart illustrating a procedure of a method for designing a control logic of a compression ignition engine.

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

  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. In FIG. 1, the intake side is the left side of the drawing, and the exhaust side is the right side of the drawing. 2 and 3, the intake side is the right side of the drawing, and the exhaust side is the left side of the drawing. FIG. 4 is a block diagram illustrating the configuration of the engine control apparatus.

  The engine 1 is a four-stroke engine that operates when the combustion chamber 17 repeats an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke. The engine 1 is mounted on a four-wheeled vehicle. The vehicle travels when the engine 1 is driven. The fuel of the engine 1 is gasoline in this configuration example. The fuel may be a liquid fuel containing at least gasoline. The fuel may be gasoline containing bioethanol, for example.

(Engine configuration)
The engine 1 includes a cylinder block 12 and a cylinder head 13 placed on the cylinder block 12. A plurality of cylinders 11 are formed inside the cylinder block 12. 1 and 2, only one cylinder 11 is shown. 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 defines a 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 a space formed by the piston 3, the cylinder 11, and the cylinder head 13 regardless of the position of the piston 3.

  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 as shown in the lower diagram of FIG. 2. The inclined surface 1311 has an upward slope from the intake side toward an injection axis X2 of an injector 6 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. In this configuration example, the cavity 31 has a shallow dish shape. The center of the cavity 31 is shifted to the exhaust side from the center 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 combining SI combustion and CI combustion in a part of the operation region. SPCCI combustion uses the heat generated by SI combustion and the pressure rise to control CI combustion. The engine 1 is a compression ignition engine. However, the engine 1 does not need to increase the temperature of the combustion chamber 17 (that is, the compression end temperature) when the piston 3 reaches the compression top dead center. The engine 1 can set the geometric compression ratio relatively low. Lowering the geometric compression ratio is advantageous for reducing cooling loss and mechanical loss. The geometric compression ratio of the engine 1 is 14 to 17 in the regular specification (low octane fuel having a fuel octane number of about 91), and 15 in the high-octane specification (high octane fuel having a fuel octane number of about 96). It is good also as ~ 18.

  An intake port 18 is formed in the cylinder head 13 for each cylinder 11. The intake port 18 has a first intake port 181 and a second intake port 182 as shown in FIG. The intake port 18 communicates with the combustion chamber 17. Although not shown in detail, the intake port 18 is a so-called tumble port. That is, the intake port 18 has such a shape that a tumble flow is formed in the combustion chamber 17.

  An intake valve 21 is disposed in the intake port 18. The intake valve 21 opens and closes between the combustion chamber 17 and the intake port 18. The intake valve 21 is opened and closed at a predetermined timing by a valve operating mechanism. The valve mechanism may be a variable valve mechanism that varies valve timing and / or valve lift. 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 motorized S-VT 23 continuously changes the rotation phase of the intake camshaft within a predetermined angle range. The opening timing and 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 provided 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 disposed in the exhaust port 19. The exhaust valve 22 opens and closes between the combustion chamber 17 and the exhaust port 19. The exhaust valve 22 is opened and closed at a predetermined timing by a valve mechanism. This valve mechanism may be a variable valve mechanism that makes the valve timing and / or valve lift variable. In this configuration example, as shown in FIG. 4, the variable valve mechanism has an exhaust electric S-VT 24. The exhaust electric S-VT 24 continuously changes the rotational phase of the exhaust camshaft within a predetermined angle range. The opening timing and 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 motor S-VT 23 and the exhaust motor S-VT 24 adjust the length of the overlap period in which both the intake valve 21 and the exhaust valve 22 are opened. When the length of the overlap period is increased, the residual gas in the combustion chamber 17 can be scavenged. Further, an internal EGR (Exhaust Gas Recirculation) gas can be introduced into the combustion chamber 17 by adjusting the length of the overlap period. The internal EGR system is composed of an intake electric S-VT 23 and an exhaust electric S-VT 24. Note that the internal EGR system is not necessarily configured by S-VT.

  An injector 6 is attached to the cylinder head 13 for each cylinder 11. The injector 6 directly injects fuel into the combustion chamber 17. The injector 6 is an example of a fuel injection unit. The injector 6 is disposed in a valley portion of the pent roof where the inclined surface 1311 and the inclined surface 1312 intersect. As shown in FIG. 2, the injection axis X <b> 2 of the injector 6 is located on the exhaust side with respect to the center axis X <b> 1 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 coincides with the center of the cavity 31. The injector 6 faces the cavity 31. The injection axis X2 of the injector 6 may coincide with the center axis X1 of the cylinder 11. In the case of the configuration, the injection axis X2 of the injector 6 and the center of the cavity 31 may coincide.

  Although not shown in detail, the injector 6 is constituted by a multi-injection type fuel injection valve having a plurality of injection holes. The injector 6 injects the fuel so that the fuel spray spreads radially from the center of the combustion chamber 17 as indicated by a two-dot chain line in FIG. In the present configuration example, the injector 6 has ten nozzle holes, and the nozzle 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 the crankshaft 15. The common rail 64 stores the fuel pumped from the fuel pump 65 at a high fuel pressure. When the injector 6 is opened, the fuel stored in the common rail 64 is injected into the combustion chamber 17 from the injection port of the injector 6. The fuel supply system 61 can supply high pressure fuel of 30 MPa or 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 disposed closer to the intake side than the center 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 the top to the bottom toward the center of the combustion chamber 17. As shown in FIG. 2, the electrode of the spark plug 25 faces the combustion chamber 17 and is located near the ceiling surface of the combustion chamber 17. The spark plug 25 may be disposed on the exhaust side of the center axis X1 of the cylinder 11. Further, the spark plug 25 may be disposed on the central axis X1 of the cylinder 11.

  An intake passage 40 is connected to one side 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 disposed 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 disposed between the air cleaner 41 and the surge tank 42 in the intake passage 40. The throttle valve 43 adjusts the amount of fresh air introduced into the combustion chamber 17 by adjusting the opening of the valve.

  A supercharger 44 is also arranged in the intake passage 40 downstream of the throttle valve 43. The supercharger 44 supercharges the gas introduced into the combustion chamber 17. In this configuration example, the supercharger 44 is a mechanical supercharger driven by the engine 1. The mechanical supercharger 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 between the supercharger 44 and the engine 1 or interrupts the transmission of the driving force. As will be described later, when the ECU 10 switches between disconnection and connection of the electromagnetic clutch 45, the supercharger 44 is switched between on and off.

  An intercooler 46 is disposed downstream of the supercharger 44 in the intake passage 40. The intercooler 46 cools the gas compressed in the supercharger 44. The intercooler 46 may be configured to be, for example, a water cooling type or an oil cooling 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 so as to bypass the supercharger 44 and the intercooler 46. An air bypass valve 48 is disposed in the bypass passage 47. The air bypass valve 48 adjusts the flow rate of the gas flowing through the bypass passage 47.

  The ECU 10 fully opens the air bypass valve 48 when the supercharger 44 is turned off (that is, when the electromagnetic clutch 45 is disconnected). 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 natural intake state.

  When the supercharger 44 is turned on, the engine 1 operates in a supercharged state. The ECU 10 adjusts the opening degree of the air bypass valve 48 when the supercharger 44 is turned on (that is, when the electromagnetic clutch 45 is connected). A part of the gas that has passed through the supercharger 44 flows back upstream of the supercharger 44 through the bypass passage 47. When the ECU 10 adjusts the opening degree of the air bypass valve 48, the supercharging pressure of the gas introduced into the combustion chamber 17 changes. The supercharging is defined as the time when the pressure in the surge tank 42 exceeds atmospheric pressure, and the non-supercharging is defined as the time when the pressure in the surge tank 42 is lower than atmospheric pressure. Also good.

  In this configuration example, the supercharger 44, the bypass passage 47, and the air bypass valve 48 constitute a supercharging system 49.

  The engine 1 has a swirl generator 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 disposed 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 adjustment valve that can narrow the cross section of the secondary passage 402. When the opening of the swirl control valve 56 is small, the intake flow rate flowing into the combustion chamber 17 from the first intake port 181 is relatively large, and the intake flow rate flowing into the combustion chamber 17 from the second intake port 182 is relatively large. Since there are few, the swirl flow in the combustion chamber 17 becomes strong. When the opening of the swirl control valve 56 is large, the intake flow rate flowing into the combustion chamber 17 from each of the first intake port 181 and the second intake port 182 becomes substantially uniform, so 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 circulates in the counterclockwise direction in FIG. 3 as indicated by the white arrow (see also the white arrow in FIG. 2).

  An exhaust passage 50 is connected to the other side 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 exhaust gas discharged from the combustion chamber 17 flows. Although the detailed illustration is omitted, the upstream portion of the exhaust passage 50 constitutes an independent passage branched 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 disposed in the exhaust passage 50. Although not shown, the upstream catalytic converter is disposed in the engine room. The upstream catalytic converter includes a three-way catalyst 511 and a GPF (Gasoline Particulate Filter) 512. The downstream catalytic converter is disposed 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 one having a three-way catalyst. Furthermore, the arrangement 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. A downstream end of the EGR passage 52 is connected to an upstream portion of the supercharger 44 in the intake passage 40. The EGR gas flowing through the EGR passage 52 does not pass through the air bypass valve 48 of the bypass passage 47 and enters the upstream portion of the supercharger 44 in the intake passage 40.

  A water-cooled EGR cooler 53 is disposed in the EGR passage 52. The EGR cooler 53 cools the exhaust gas. An EGR valve 54 is also disposed in the EGR passage 52. The EGR valve 54 adjusts 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 includes an external EGR system and an internal EGR system. The external EGR system can supply exhaust gas having a temperature lower than that of the internal EGR system to the combustion chamber 17.

  The control device for the compression ignition 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 ROM A memory 102 configured by (Read Only Memory) and storing programs and data, and an input / output bus 103 for inputting and outputting electrical signals are provided. The ECU 10 is an example of a control unit.

  As shown in FIGS. 1 and 4, various sensors SW <b> 1 to SW <b> 17 are connected to the ECU 10. Sensors SW1-SW17 output a signal to ECU10. The sensors include the following sensors.

Air flow sensor SW1: The air flow sensor SW2 is disposed downstream of the air cleaner 41 in the intake passage 40 and measures the flow rate of fresh air flowing through the intake passage 40. The first intake air temperature sensor SW2 is disposed downstream of the air cleaner 41 in the intake passage 40 and , Measuring the temperature of fresh air flowing through the intake passage 40. First pressure sensor SW3: disposed downstream of the connection position of the EGR passage 52 in the intake passage 40 and upstream of the supercharger 44, and the supercharger 44. The second intake air temperature sensor SW4 is disposed downstream of the supercharger 44 in the intake passage 40 and upstream of the connection position of the bypass passage 47 and flows out of the supercharger 44. Second pressure sensor SW5 for measuring the temperature of the gas: It is attached to the surge tank 42 and measures the pressure of the gas downstream of the supercharger 44. Finger pressure sensor SW6: attached to the cylinder head 13 corresponding to each cylinder 11, and measures the pressure in each combustion chamber 17. Exhaust temperature sensor SW7: exhaust gas disposed in the exhaust passage 50 and exhausted from the combustion chamber 17 Measuring the temperature of the gas Linear O ₂ sensor SW8: Arranged upstream of the upstream catalytic converter in the exhaust passage 50 and measuring the oxygen concentration in the exhaust gas Lambda O ₂ sensor SW9: Three-way in the upstream catalytic converter Water temperature sensor SW10, which is disposed downstream of the catalyst 511 and measures the oxygen concentration in the exhaust gas, is attached to the engine 1, and the temperature of the cooling water is measured. Crank angle sensor SW11: which is attached to the engine 1, and the crankshaft Measures 15 rotation angle Accelerator opening sensor SW12: Accelerator pedal machine The intake cam angle sensor SW13 is attached to the engine 1 and measures the rotation angle of the intake camshaft. The exhaust cam angle sensor SW14 is attached to the engine 1. EGR differential pressure sensor SW15 is mounted and measures the differential angle upstream and downstream of the EGR valve 54 and is measured in the EGR passage 52. Fuel pressure sensor SW16: Common rail of the fuel supply system 61 64, and measures the pressure of fuel supplied to the injector 6. Third intake temperature sensor SW17: attached to the surge tank 42 and introduced into the combustion chamber 17, in other words, the temperature of the gas in the surge tank 42 Measure the intake air temperature.

  The ECU 10 determines the operating state of the engine 1 based on the signals of these 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 sends an electric signal related to the calculated control amount to the injector 6, spark plug 25, intake electric S-VT 23, exhaust electric S-VT 24, fuel supply system 61, throttle valve 43, EGR valve 54, supercharger 44. Output 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 from 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 supercharging pressure and the differential pressure across the supercharger 44 obtained from the signals of the first pressure sensor SW3 and the second pressure sensor SW5. By performing feedback control, the supercharging pressure becomes the target supercharging pressure.

  Further, the ECU 10 sets a target EGR rate (that is, a ratio of EGR gas to all gases 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 differential pressure across the EGR valve 54 obtained from the signal of the EGR differential pressure sensor SW15. By performing feedback control that adjusts the opening degree of the EGR valve 54 based on the above, the amount of external EGR gas introduced into the combustion chamber 17 becomes the target EGR gas amount.

Further, the ECU 10 executes air-fuel ratio feedback control when a predetermined control condition is satisfied. Specifically, the ECU 10 performs the fuel injection of the injector 6 so that the air-fuel ratio of the air-fuel mixture becomes a desired value based on the oxygen concentration in the exhaust gas measured by the linear O ₂ sensor SW8 and the lambda O ₂ sensor SW9. Adjust the amount.

  Details of other controls of the engine 1 by the ECU 10 will be described later.

(SPCCI combustion concept)
The engine 1 performs combustion by compression self-ignition in a predetermined operation state mainly for the purpose of improving fuel consumption and exhaust gas performance. In the combustion by self-ignition, when the temperature in the combustion chamber 17 before the start of compression varies, the timing of self-ignition greatly changes. Therefore, the engine 1 performs SPCCI combustion combining SI combustion and CI combustion.

  In the SPCCI combustion, the spark 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 in the combustion chamber 17 generates heat from the SI combustion. 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 increases due to flame propagation.

  By adjusting the calorific value of the SI combustion, it is possible to absorb the temperature variation 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 a target timing.

  In SPCCI combustion, heat generation during SI combustion is milder than heat generation during CI combustion. In the waveform of the heat generation rate in the SPCCI combustion, as illustrated in FIG. 5, the rising slope is smaller than the rising slope in the CI combustion waveform. The pressure fluctuation (dp / dθ) in the combustion chamber 17 is also gentler during SI combustion than during CI combustion.

  When the unburned mixture self-ignites after the start of SI combustion, the slope of the heat generation rate waveform may change from small to large at the timing of self-ignition. The waveform of the heat release 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 compression top dead center, it is avoided that the inclination of the waveform of the heat release rate becomes too large. Pressure fluctuation (dp / dθ) during CI combustion is also relatively gentle.

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

  When CI combustion ends, SPCCI combustion ends. CI combustion has a shorter combustion period than SI combustion. In SPCCI combustion, the combustion end timing is earlier than SI combustion.

The heat-release-rate waveform of SPCCI combustion, a first heat generation rate portion Q _SI formed by SI combustion, and the second heat generating section Q _CI formed by CI combustion, but formed continuous in this order Has been.

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 a ratio of the amount of heat generated by two combustions having different combustion forms. When the SI rate is high, the SI combustion rate is high, and when the SI rate is low, the CI combustion rate 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 ).

  The engine 1 generates a strong swirl flow in the combustion chamber 17 when performing SPCCI combustion. The strong swirl flow may be defined as a flow having a swirl ratio of 4 or more, for example. The swirl ratio can be defined as a value obtained by dividing the intake flow lateral angular velocity measured and integrated for each valve lift by the engine angular velocity. Although not shown, the intake flow lateral angular velocity can be obtained based on measurement using a known rig testing apparatus.

  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 at the central portion becomes relatively weak. Due to the eddy current caused by the velocity gradient at the boundary between the central portion and the outer peripheral portion, the turbulent energy is increased in the central portion. When the spark plug 25 ignites the air-fuel mixture in the center, the combustion speed of SI combustion increases due to high turbulent energy.

  The SI combustion flame rides on the strong swirl flow in the combustion chamber 17 and propagates in the circumferential direction. In 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 before the start of CI combustion. Generation of combustion noise can be suppressed, and torque variation between cycles can be suppressed.

(Engine operating range)
6 and 7 exemplify maps 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 in accordance with 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 map when the engine 1 is warm. The second map 502 is a map when the engine 1 is half warmed up. The third map 503 is a map when the engine 1 is cold.

  Each map 501, 502, 503 is defined by the load and the rotational speed of the engine 1. The first map 501 is roughly divided into three regions for the load level and the rotational speed level. Specifically, the three regions include idle operation and the low load region A1, the medium and high load regions A2, A3, A4, and the low load regions that are higher in the low load region A1 and the low load region A1. This is a high rotation area A5 having a higher rotational speed than the load area A1, the medium and high load areas A2, A3, and A4. The medium and high load areas A2, A3 and A4 are also composed of a medium load area A2, a high load medium rotation area A3 having a higher load than the medium load area A2, and a high load low rotation having a lower rotational speed than the high load medium rotation area A3. Divided into area A4.

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

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

  Here, the low rotation region, the medium rotation region, and the high rotation region are each when the entire operation region of the engine 1 is divided into approximately three equal parts of the low rotation region, the medium rotation region, and the high rotation region in the rotation speed direction. The low rotation region, the medium rotation region, and the high rotation region may be used. In the example of FIGS. 6 and 7, the rotation speed less than N1 is low, the rotation speed N2 or more is high rotation, and the rotation speed N1 or more and less than N2 is medium rotation. For example, the rotational speed N1 may be about 1200 rpm, and the rotational speed N2 may be about 4000 rpm, for example.

  Further, the low load region may be a region including a light load operation state, the high load region may be a region including a fully open load operation state, and the medium load may be a region between the low load region and the high load region. In addition, the low load region, the medium load region, and the high load region are respectively obtained when the entire operation region of the engine 1 is divided into approximately three equal parts of the low load region, the medium load region, and the high load region. A low load region, a medium load region, and a high load region may be used.

  Maps 501, 502, and 503 in FIG. 6 respectively show the state of the air-fuel mixture and the combustion mode in each region. A map 504 in FIG. 7 corresponds to the first map 501, and in this map, the state of the air-fuel mixture and the combustion mode in each region, the opening of the swirl control valve 56 in each region, and the drive region of the supercharger 44 And a non-driving region. The engine 1 includes 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 in FIG. 7 will be described in detail with reference to the fuel injection timing and the ignition timing shown in FIG. The horizontal axis in FIG. 8 is the crank angle. Note that reference numerals 601, 602, 603, 604, 605, and 606 in FIG. 8 correspond to the operating state of the engine 1 indicated by reference numerals 601, 602, 603, 604, 605, and 606 in the map 504 in 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 fuel injection timing (reference numerals 6011 and 6012) and ignition timing (reference numeral 6013) when the engine 1 is operating in the operation state 601 in the low load region A1, and a combustion waveform (that is, A waveform showing a 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 numerals 6021 and 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. , 603 indicates the fuel injection timing (reference numerals 6031, 6032) and 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. Show. The operating states 601, 602, and 603 have the same engine 1 rotational speed and different loads. The operation state 601 has the lowest load (ie, light load), the operation state 602 has the next lowest load (ie, low load), and the operation state 603 has the highest load.

  In order to improve the fuel efficiency performance of the engine 1, the EGR system 55 introduces EGR gas into the combustion chamber 17. Specifically, the intake electric S-VT 23 and the exhaust electric S-VT 24 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. 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 hot exhaust gas is introduced into the combustion chamber 17, the temperature in the combustion chamber 17 increases. This is advantageous for stabilizing the SPCCI combustion. The intake electric S-VT 23 and the exhaust electric S-VT 24 may be provided with a negative overlap period in which both the intake valve 21 and the exhaust valve 22 are closed.

  The swirl generator forms a strong swirl flow in the combustion chamber 17. The swirl ratio is, for example, 4 or more. The swirl control valve 56 has a predetermined opening 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 the fuel into the combustion chamber 17 a plurality of times during the intake stroke (reference numerals 6011, 6012, 6021, 6022, 6031, 6032). The air-fuel mixture is stratified by the multiple fuel injections and the 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 part is 20 or more and 30 or less, and the A / F of the air-fuel mixture in the outer peripheral part is 35 or more. Note that 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, generation of RawNOx during SI combustion can be suppressed. Moreover, CI combustion is stabilized by making A / F of the air-fuel mixture in the outer peripheral portion 35 or more.

  The air-fuel ratio (A / F) of the air-fuel mixture is leaner than the stoichiometric 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 carrying out like this, generation | occurrence | production of RawNOx can be suppressed and exhaust gas performance can be improved.

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

  In the operating state 602 where the load of the engine 1 is high, 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 engine 1 is in a higher operating state 603, the injector 6 starts the second injection 6032 performed in the latter half of the intake stroke at a timing that is 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 6: 4, for example.

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

  As described above, since the air-fuel mixture in the center has a relatively high fuel concentration, the ignitability is improved and SI combustion by flame propagation is stabilized. By stabilizing the SI combustion, the CI combustion starts at an appropriate timing. In SPCCI combustion, controllability of CI combustion is improved. Generation of combustion noise is suppressed. Further, by performing SPCCI combustion with the A / F of the air-fuel mixture leaner than the stoichiometric air-fuel ratio, the fuel efficiency performance of the engine 1 can be greatly improved. The low load area A1 corresponds to layer 3 described later. Layer 3 extends to the light load operation region and includes the minimum load operation state.

(Medium and high load area)
Even when the engine 1 is operating in the middle and high load region, the engine 1 performs SPCCI combustion similarly to the low load region.

  Reference numeral 604 in FIG. 8 denotes a fuel injection timing (reference numerals 6041 and 6042) and an ignition timing (reference numeral 6043) when the engine 1 is operating in the operation state 604 in the medium load area A2, and The combustion waveform (reference numeral 6044) is shown. Reference numeral 605 indicates a fuel injection timing (reference numeral 6051) and 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-VT 23 and the exhaust electric S-VT 24 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. 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, an external EGR gas having a temperature lower than that of the internal EGR gas is introduced into the combustion chamber 17. The external EGR gas adjusts the temperature in 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 make the EGR gas including the internal EGR gas and the external EGR gas zero at a full open load.

  Moreover, in the middle / high load region A2 and the high load / medium rotation region A3, the swirl control valve 56 has a predetermined opening 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, the swirl control valve 56 is open in the high load low rotation region A4.

  The air-fuel ratio (A / F) of the air-fuel mixture is the stoichiometric air-fuel ratio (A / F≈14.7) in the entire combustion chamber 17. As the three-way catalysts 511 and 513 purify the exhaust gas discharged from the combustion chamber 17, the exhaust gas performance of the engine 1 is improved. The A / F of the air-fuel mixture may be set within 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 mid-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 made richer (that is, the excess air ratio λ of the air-fuel mixture is λ ≦ 1).

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

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

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

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

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

(Supercharger operation)
Here, as shown in a map 504 in FIG. 7, the turbocharger 44 is off in a part of the low load area A1 and a part of the medium / high load area A2 (see S / C OFF). Specifically, the supercharger 44 is off in the low rotation side region in the low load region A1. In the region on the high rotation side in the low load region A1, the supercharger 44 is on in order to ensure a necessary intake charge amount corresponding to the increase in the rotational speed of the engine 1. Further, the supercharger 44 is off in a partial region on the low load / low rotation side in the medium / high load region A2. In the region on the high load side in the middle and high load region A2, the supercharger 44 is on in order to ensure a necessary intake charge amount corresponding to the increase in the fuel injection amount. The supercharger 44 is also on in the high rotation side region in the medium and high load region A2.

  In each of the high load mid-rotation region A3, the high load low rotation region A4, and the high rotation region A5, the supercharger 44 is on (see S / C ON).

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

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

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

  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 make the EGR gas zero at full load.

  The swirl control valve 56 is fully open. A swirl flow is not generated in the combustion chamber 17 and only a tumble flow is generated. By fully opening the swirl control valve 56, the charging 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 stoichiometric air-fuel ratio (A / F≈14.7) in the entire combustion chamber 17. The excess air ratio λ of the air-fuel mixture may be 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 in a lump (reference numeral 6061). By starting fuel injection during the intake stroke, a homogeneous or substantially homogeneous mixture is formed in the combustion chamber 17. Moreover, since the fuel vaporization time can be secured for a long time, unburned loss can be reduced.

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

(Map layer structure)
As shown in FIG. 9, the maps 501, 502, and 503 of the engine 1 illustrated in FIG. 6 are configured by combining three layers of layer 1, layer 2, and layer 3.

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

  Layer 2 is a layer overlying layer 1. Layer 2 corresponds to a part of the operating region of 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 overlying layer 2. Layer 3 corresponds to the low load area A1 of the first map 501.

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

  The wall temperature of the combustion chamber 17 is less than a first predetermined wall temperature and a second predetermined wall temperature (for example, 30 ° C.) or more, and the intake air temperature is less than the first predetermined intake temperature and the second predetermined intake temperature (for example, 25 ° C.) or more. Sometimes layer 1 and layer 2 are selected. A second map 502 is configured by overlapping these layer 1 and layer 2. In the low and medium rotation areas B1, B2, and B3 in the second map 502, the highest layer 2 is effective, and in the high rotation area B4, layer 1 is effective.

  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 the layer 1 is selected and the third map 503 is configured.

  Note that the wall temperature of the combustion chamber 17 may be substituted by the coolant temperature of the engine 1 measured by the water temperature sensor SW10, for example. Further, the wall temperature of the combustion chamber 17 may be estimated based on the temperature of the cooling water and other measurement results. The intake air temperature can be measured by, for example, the third intake air temperature sensor SW17 that measures the temperature in 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, the SPCCI combustion is performed by generating a strong swirl flow in the combustion chamber 17. In SI combustion, a flame propagates along the wall of the combustion chamber 17, so that the flame propagation of SI combustion is affected by the wall temperature. When the wall temperature is low, the flame of SI combustion is cooled, and the timing of compression ignition is delayed.

  The CI combustion in the SPCCI combustion is performed from the outer peripheral portion of the combustion chamber 17 to the central portion, and therefore is affected by the temperature of the central portion of the combustion chamber 17. If the temperature at the center is low, CI combustion becomes unstable. The temperature at the center 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 at the center of the combustion chamber 17 is high, and when the intake air temperature is low, the temperature at the center 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 the layer 1 that performs 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 air temperature, the air-fuel mixture having a substantially theoretical air-fuel ratio (that is, λ≈1) can be stably subjected to SPCCI combustion. it can. Therefore, in addition to layer 1, layer 2 is selected, and ECU 10 operates engine 1 based on second map 502. The engine 1 performs SPCCI combustion in a part of the operation region, so that 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 air temperature, the air-fuel mixture leaner than the stoichiometric air-fuel ratio can be stably subjected to SPCCI combustion. Therefore, layer 3 is selected in addition to layer 1 and layer 2, and ECU 10 operates engine 1 based on first map 501. The engine 1 causes the lean air-fuel mixture to undergo SPCCI combustion in a part of the operation region, thereby further improving the fuel efficiency performance of the engine 1.

  Next, a control example related to map layer selection 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 signals from the sensors SW1 to SW17. In subsequent step S2, ECU 10 determines whether the wall temperature of combustion chamber 17 is 30 ° C. or higher and the intake air temperature is 25 ° C. or higher. When the determination at step S2 is YES, the process proceeds to step S3, and when the determination is NO, the process proceeds to step S5. The ECU 10 selects only the 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 the wall temperature of the combustion chamber 17 is 80 ° C. or higher and the intake air temperature is 50 ° C. or higher. When the determination in step S3 is YES, the process proceeds to step S4. When the determination is 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-VT 23 according to the control logic set for the layer 2. The upper diagram in FIG. 11 (that is, the graph 1101) shows the change (horizontal axis) of the valve opening timing IVO of the intake valve 21 with respect to the load of the engine 1 (horizontal axis). The solid line corresponds to the first rotation speed when the engine 1 is relatively low, and the broken line corresponds to the second rotation speed when the engine 1 is relatively high (first rotation speed <second rotation speed). To do.

  The lower diagram of FIG. 11 (that is, the graph 1102) shows the change (vertical axis) of the valve opening timing IVO of the intake valve 21 with respect to the rotational speed of the engine 1 (horizontal axis). The solid line corresponds to the first load when the load on the engine 1 is relatively low, and the broken line corresponds to the second load when the load on 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 is advanced as it goes upward, and the positive overlap period in which both the intake valve 21 and the exhaust valve 22 are opened becomes longer. Therefore, the amount of EGR gas introduced into the combustion chamber 17 increases.

  In the layer 2, the engine 1 is operated with the A / F of the air-fuel mixture set to the stoichiometric air fuel ratio or substantially the stoichiometric air fuel ratio, and the G / F is leaner than the stoichiometric air fuel ratio. When the load on 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 on the engine 1 increases, the amount of fuel supply increases, so that combustion stability increases. The ECU 10 sets the valve opening timing of the intake valve 21 to the advance side timing. 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 on the engine 1 is further increased, the temperature in the combustion chamber 17 is further increased. The amount of internal EGR gas is reduced and the amount of external EGR gas is increased so that the temperature in the combustion chamber 17 does not become too high. For this purpose, the ECU 10 sets the valve opening timing of the intake valve 21 to the retard side timing again.

  When the load on the engine 1 is further increased and the supercharger 44 performs supercharging, the ECU 10 sets the valve opening timing of the intake valve 21 to the advance side timing again. Since a positive overlap period for opening both the intake valve 21 and the exhaust valve 22 is provided, the residual gas in the combustion chamber 17 can be scavenged.

  Note that the tendency of the opening timing of the intake valve 21 to change when the engine 1 is high and low is almost the same.

  As shown in the 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.

  When the rotational 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 advance side timing.

  When the rotational speed of the engine 1 further increases, the ECU 10 sets the valve opening timing of the intake valve 21 to a retarded timing corresponding to the rotational speed of the engine 1. Thereby, the amount of gas introduced into the combustion chamber 17 is maximized.

  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-VT 23 and the exhaust electric S-VT 24 according to the control logic set for the layer 3. And, an example of a change in the overlap period O / L of the intake valve 21 and the exhaust valve 22 is shown.

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

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

  The lower diagram of FIG. 12 (that is, the 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 load level (horizontal axis) of the engine 1. The solid line corresponds to the case where the engine 1 has the third rotational speed, and the broken line corresponds to the case where the engine 1 has the fourth rotational speed.

  In the layer 3, the engine 1 is operated by subjecting the air-fuel mixture whose A / F is leaner than the theoretical air-fuel ratio to SPCCI combustion. When the load on 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 the graph 1201, the ECU 10 sets the valve opening timing IVO of the intake valve 21 to a timing that is retarded from the exhaust top dead center. The closing timing of the intake valve 21 is a so-called delayed closing after the intake bottom dead center. Further, when the load on the engine 1 is low, the ECU 10 limits the amount of internal EGR gas introduced into the combustion chamber 17. As shown in the graph 1202, the ECU 10 sets the valve closing timing EVC of the exhaust valve 22 to a timing on the advance side. The valve closing timing EVC of the exhaust valve 22 approaches the exhaust top dead center.

  When the load on the engine 1 increases, the amount of fuel supply increases, so the ECU 10 does not limit the amount of gas introduced into the combustion chamber 17. Further, the ECU 10 increases the amount of internal EGR gas introduced into the combustion chamber 17 in order to stabilize the SPCCI combustion of the air-fuel mixture leaner than the stoichiometric air-fuel ratio. The ECU 10 sets the valve opening timing IVO of the intake valve 21 to a timing on the more advanced side than the exhaust top dead center. Further, the ECU 10 sets the valve closing timing EVC of the exhaust valve 22 to a timing that is retarded from the exhaust top dead center. As a result, as shown in the graph 1203, when the load on the engine 1 is increased, the overlap period in which both the intake valve 21 and the exhaust valve 22 are opened becomes longer.

  When the load on the engine 1 is further increased, the ECU 10 reduces the amount of 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 close to the exhaust top dead center. The overlap period between the intake valve 21 and the exhaust valve 22 is shortened. Further, the ECU 10 sets the valve opening timing of the intake valve 21 to the retarded side timing when the load of the engine 1 is high and the rotation speed of the engine 1 is high than when the rotation speed is low. The amount of gas introduced into the combustion chamber 17 is maximized.

  Note that, in the low load region surrounded by the one-dot chain line in FIG. 12, the engine 1 may perform a reduced-cylinder operation for the purpose of improving the fuel efficiency. When performing the reduced cylinder operation, the amount of gas introduced into the combustion chamber 17 and the amount of internal EGR gas are not limited. The ECU 10 sets the valve opening timing of the intake valve 21 to the advance side timing and the valve closing timing of the exhaust valve 22 to the retard side timing as shown by the two-dot chain lines in the graphs 1201 and 1202. 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 combusts so that the combustion in the combustion chamber 17 becomes the combustion of the SI rate corresponding to the operating state. Calculations for adjusting the state quantity in the chamber 17, adjusting the injection amount, adjusting the injection timing, and adjusting the ignition timing are performed.

  First, in step S131, the ECU 10 reads signals from the sensors SW1 to SW17. 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 a target SI rate (that is, a target heat amount ratio). The target SI rate is set according to the operating state of the engine 1.

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

  Returning to the flow of FIG. 13, in the subsequent step S <b> 133, the ECU 10 sets a target in-cylinder state quantity 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 requires 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 quantity. In addition, the phase angle of the intake motor S-VT 23 and the exhaust motor S-VT 24 (that is, the valve timing of the intake valve 21 and the valve timing of the exhaust valve 22) is set. The ECU 10 sets control amounts of these devices based on a 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-VT 23 and the exhaust electric S-VT 24 based on the set control amount. . As each device operates based on the signal from 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 quantity in the combustion chamber 17 based on the set control amounts of the respective devices. The state quantity predicted value is a value obtained by predicting 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 obtained by estimating 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 in the compression stroke and setting the ignition timing, 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 amount predicted value. When performing divided injection during the intake stroke, the injection amount of each injection is set. When 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 an injection timing based on a preset map.

  In step S139, the ECU 10 sets the ignition timing based on the state quantity estimated value and the fuel injection result during the compression stroke. In step S1310, 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 set ignition timing.

  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 pressure change in the combustion chamber 17 measured by the finger pressure 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 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 estimation in step S134 in the subsequent cycles. The ECU 10 opens 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-VT 23 and the exhaust electric S-VT 24 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.

  The ECU 10 also performs injection in the compression stroke in step S138 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 state quantity estimated value. The timing is advanced from the injection timing based on the map. On the other hand, when the ECU 10 predicts that the temperature in the combustion chamber 17 becomes higher than the target temperature based on the state quantity estimated value, in step S138, the ECU 10 is in the compression stroke so that the ignition timing can be retarded. The injection timing is retarded from the injection timing based on the map.

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

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 the start of SI combustion is accelerated, sufficient heat generation is possible by SI combustion, and therefore, the self-ignition timing θ _{CI of the} unburned mixture is prevented 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.

  If the temperature in the combustion chamber 17 is high, the unburned mixture self-ignites immediately after the start of SI combustion by spark ignition, and the SI rate deviates from the target SI rate. In this case, 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 self-ignition timing θ _{CI of the} unburned 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 motor S-VT 23, and an exhaust motor S-VT 24. It is configured to 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 ignition timing. By adjusting the injection timing and the ignition timing, for example, the difference between cylinders can be corrected, or the self-ignition timing can be finely adjusted. By performing the adjustment of the SI rate in two stages, the engine 1 can accurately realize the target SPCCI combustion corresponding to the operation 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 on the engine 1 is high, the ECU 10 increases the target SI rate than when the engine 1 is low. Therefore, the ECU 10 has a higher SI rate when the load on the engine 1 is high than when the load is low. The signal is output to at least 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.

(Engine control logic design method)
When designing the control logic of the engine 1 that executes the above-described SPCCI combustion, parameters related to the control amount of each device are set. For example, in the case of the spark plug 25, ignition energy and ignition timing corresponding to the operating state of the engine 1 are set. In the case of the intake electric S-VT 23, the valve timing of the intake valve 21 corresponding to the operating state of the engine 1 is set. Hereinafter, setting of the valve timing 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 performs SPCCI combustion, in order to suppress combustion noise and realize stable SPCCI combustion, the temperature in the combustion chamber 17 at the timing at which CI combustion starts (θ _CI : see FIG. 5) is appropriately set. The inventors of the present invention have found that it is necessary to adjust to a suitable temperature. That is, when the temperature in the combustion chamber 17 is low, the ignitability of CI combustion is reduced. When the temperature in the combustion chamber 17 is high, combustion noise increases.

The temperature in the combustion chamber 17 at the timing θ _{CI at} which CI combustion starts 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 newly found out that an appropriate IVC range exists in the range of the geometric compression ratio ε at which SPCCI combustion can occur in order to put an engine that performs 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 an engine that performs a specific combustion mode called SPCCI combustion into practical use. There is a novelty in the point that I found out. Further, a relational expression between the geometric compression ratio ε and the valve closing timing IVC of the intake valve 21 described later is also new.

  Furthermore, 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 novelty in that 21 valve closing timing IVC is set.

  The engine 1 performs the SPCCI combustion (layer 2) when the A / F of the air-fuel mixture is made richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio and the G / F is leaner than the stoichiometric air-fuel ratio (layer 2). Switching between when A / F is leaner than the stoichiometric air-fuel ratio and SPCCI combustion is performed (layer 3). 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 in layer 3 and the valve closing timing IVC of the intake valve 21.

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

  The range in which SPCCI combustion is established is indicated by the hatched range in FIG. The range in which SPCCI combustion is established is sandwiched between the “advance limit” line and the “retard angle limit” line. The “advance limit” line means that abnormal combustion occurs when the combustion center of gravity advances from this line, and therefore, SPCCI combustion is not established. Further, the “retard angle limit” line means that if the combustion center of gravity is retarded from this line, self-ignition does not occur, and therefore SPCCI combustion is not established.

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

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

  When the load of the engine 1 is low (“light load”), the range in which SPCCI combustion is established is on the advance side, so the EGR rate is adjusted to some extent as shown by the double-ended arrows in FIG. At the same time, by adjusting the combustion center of gravity to the advance side and the retard side, SPCCI combustion corresponding to MBT can be realized.

  When the load on the engine 1 increases, the amount of air introduced into the combustion chamber 17 must be increased in response to an 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 be introduced into the combustion chamber 17. However, it is difficult to introduce a large amount of both air and EGR gas into the combustion chamber 17 due to the limit of the supercharging capability of the supercharger 44. Therefore, when the load on the engine 1 increases, the range in which SPCCI combustion is established becomes the retarded angle side. When the combustion center of gravity of SPCCI combustion is to be brought closest to MBT when the load of engine 1 is high, SPCCI combustion must be performed at point Y in FIG. Point Y corresponds to the operating state of engine 1 at the upper limit load at which the air-fuel mixture with A / F at the stoichiometric air-fuel ratio can be subjected to SPCCI combustion. 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 or adjusting the combustion center of gravity.

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

In order to know the temperature in 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 obtained from the engine 1. Decided to do. Specifically, the inventors of the present application measure various parameters when the engine 1 is operating at the upper limit load in the layer 2, and estimate the actual temperature in the combustion chamber 17 at θ _CI from the measured parameters. did. The inventors of the present application determined an average value of a plurality of estimated temperatures as a reference temperature Tth1. If the reference temperature Tth1 temperature of the combustion chamber 17 in the theta _CI, it is possible to realize a SPCCI combustion in the layer 2.

Here, SPCCI combustion, as described above, by adjusting the ignition timing, it is possible to adjust the theta _CI. However, when the engine 1 is operating at the upper limit load in the layer 2 and the temperature in the combustion chamber 17 at θ _CI exceeds the reference temperature Tth1, the θ _CI is adjusted even if the ignition timing is adjusted. Can not do. On the other hand, the temperature in the combustion chamber 17 in the theta _CI is, if the reference temperature Tth1 or less, by adjusting the ignition timing (specifically advances the ignition timing), the temperature in the combustion chamber 17 in theta _CI Can be increased to the reference temperature Tth1, and SPCCI combustion can be realized.

Therefore, in order to realize SPCCI combustion in the layer 2, the temperature of the combustion chamber 17 in the theta _CI is such as not to exceed the reference temperature Tth1, the relationship between ε and IVC are required.

Therefore, as conceptually shown in FIG. 16, the inventors of the present invention have a relationship between ε and IVC in a matrix composed of two parameters, a geometric compression ratio ε and a valve closing timing IVC of the intake valve 21. While changing the value for each, the temperature of the combustion chamber 17 at the time of θ _CI was estimated using the model of the engine 1. A combination of ε and IVC that is equal to or lower than the reference temperature Tth1 can realize SPCCI combustion in the layer 2. As shown in FIG. 16, when the geometric compression ratio ε is high and the 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. End up.

FIG. 16 is a matrix in which IVC is set after the intake bottom dead center. Although not shown, the present inventors, also the matrix set the IVC previously intake bottom dead center, by performing the estimation calculation of the temperature of the theta _CI when the combustion chamber 17, and the reference temperature Tth1 or less A combination of ε and IVC was obtained.

  The upper graph 1701 in FIG. 17 shows approximate expressions (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 a combination of ε and IVC that are equal to or lower than the reference temperature Tth1 on the plane of the graph 1701, and based on the plotted points, approximate expression (I) ( II) has been determined.

A graph 1701 corresponds to the case where the rotational speed of the engine 1 is 2000 rpm. The approximate expressions (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
It is.

  In the graph 1701, the combination of ε and IVC on the left side of the approximate expressions (I) and (II) and ε = 17, the temperature of the combustion chamber 17 during CI combustion is equal to or lower than the reference temperature Tth1. This combination enables SPCCI combustion of an air-fuel mixture in which A / F is the stoichiometric air-fuel ratio and G / F is 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 the layer 2.

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

  The temperature of the combustion chamber 17 in which the SPCCI combustion is performed is a result of two pressure increases, that is, a pressure increase due to the compression work of the piston 3 in the compression stroke and a pressure increase resulting from the heat generated by the 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 increase due to the compression work of the piston 3 is small. If the flame propagation in the SPCCI combustion advances and the pressure rise caused by the heat generated by the SI combustion does not increase considerably, the in-cylinder temperature cannot be increased to the ignition temperature. As a result, the amount of air-fuel mixture that undergoes compression self-ignition is small, and a large amount of air-fuel mixture is combusted by flame propagation, so the combustion period is long and fuel efficiency is reduced. In order to maximize the fuel efficiency by causing the CI combustion stably in the SPCCI combustion, it is necessary to keep the effective compression ratio above a certain value. Therefore, the relationship between ε and IVC must be determined.

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

If time, the temperature in the combustion chamber 17 in the theta _CI reference temperature Tth2 or the engine 1 is operating at light load, by delaying the ignition timing, it is possible to realize a SPCCI combustion. However, when the temperature of the combustion chamber 17 in the theta _CI is lower than the reference temperature Tth2, the temperature is too low, be advanced ignition timing, it is impossible to realize a SPCCI combustion.

Therefore, in order to realize SPCCI combustion in the layer 2, the temperature of the combustion chamber 17 in the theta _CI is such that the reference temperature Tth2 or more, the relationship between ε and IVC are required.

  As in the matrix shown in FIG. 16, the inventors of the present application have values for each of ε and IVC in a matrix composed of two parameters of the geometric compression ratio ε and the closing timing IVC of the intake valve 21. The temperature of the combustion chamber 17 during CI combustion was estimated using the model of the engine 1 while changing the engine. In this matrix, the combination of ε and IVC that is equal to or higher than the reference temperature Tth2 can realize SPCCI combustion in the layer 2.

A graph 1701 in FIG. 17 also shows approximate expressions (III) and (IV) calculated based on a combination of ε and IVC that are equal to or higher than the reference temperature Tth2. The approximate expressions (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
It is.

  In the graph 1701, in the combination of ε and IVC on the right side of the approximate expressions (III) and (IV), the temperature of the combustion chamber 17 during CI combustion is equal to or higher than the reference temperature Tth2. This combination enables SPCCI combustion of an air-fuel mixture in which A / F is the stoichiometric air-fuel ratio and G / F is 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 with IVC = about 20 deg. IVC = 20 deg. ABDC corresponds to the valve closing timing (that is, the best IVC) at which the amount of gas introduced into the combustion chamber 17 becomes maximum when the rotational speed of the engine 1 is 2000 rpm. Further, IVC = 120 deg. ABDC is a retard limit of the closing timing IVC of the intake valve 21, and IVC = −80 deg. ABDC is an advance limit of the closing timing IVC of the intake valve 21.

  In FIG. 17, the combination of ε and IVC within the range surrounded by the approximate expressions (I), (II), (III), and (IV) is a combination that enables practical use of the engine 1 that performs SPCCI combustion in the layer 2. It is. In other words, the combination of ε and IVC outside this range cannot put the engine 1 that performs SPCCI combustion in the layer 2 into practical use.

  When determining the valve closing timing IVC of the intake valve 21 when the engine 1 is operated in the layer 2, the designer must determine IVC within the ε-IVC establishment range indicated by the hatching in FIG.

Specifically, if the designer defines the geometric compression ratio ε as 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 determined so as to satisfy the above.

Further, when the designer defines the geometric compression ratio ε as 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 determined so as to satisfy the above.

Furthermore, when 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 determined so as to satisfy the above.

  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 theoretically SPCCI combustion of an air-fuel mixture leaner than the air-fuel ratio is realized.

  The valve closing timing IVC is set for each operation state determined by the load and the rotational speed of the engine 1 in the layer 2. The example shown by the solid line in FIG. 17 is the ε-IVC establishment range when the rotational speed of the engine 1 is 2000 rpm as described above. When the rotational speed of the engine 1 changes, the ε-IVC establishment range also changes. As the engine 1 increases in speed, the best IVC is retarded. Specifically, when the rotational speed of the engine 1 is 3000 rpm, the best IVC is about 22 deg. ABDC. As shown by a broken line in FIG. 17, the ε-IVC establishment range when the rotation speed of the engine 1 is 3000 rpm is only about 2 deg. Relative to the ε-IVC establishment range when the rotation speed of the engine 1 is 2000 rpm. Move parallel to the retard side.

Therefore, if the designer determines the geometric compression ratio ε to be 10 ≦ ε <17, when the rotational speed of the engine 1 is 3000 rpm,
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 determined so as to satisfy the above.

In addition, when the designer sets the geometric compression ratio ε to 17 ≦ ε <20, when the rotational speed of the engine 1 is 3000 rpm,
−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 determined so as to satisfy the above.

Furthermore, when 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 determined so as to satisfy the above.

  When the engine 1 has a rotational speed of 4000 rpm, the best IVC is about 28 deg. ABDC. The ε-IVC establishment range when the rotation speed of the engine 1 is 4000 rpm is about 8 deg. Min with respect to the ε-IVC establishment range when the rotation speed of the engine 1 is 2000 rpm, as shown by a one-dot chain line in FIG. Only move to the retard side.

Therefore, if the designer sets the geometric compression ratio ε to 10 ≦ ε <17, when the engine 1 has a rotational speed of 4000 rpm,
^{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 determined so as to satisfy the above.

Further, if the designer sets the geometric compression ratio ε to 17 ≦ ε <20, when the rotational speed of the engine 1 is 4000 rpm,
^{-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 determined so as to satisfy the above.

Furthermore, when 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 determined so as to satisfy the above.

The correction term C related to the rotational speed NE (rpm) of the engine 1 is
C = 3.3 × 10 ⁻¹⁰ NE ³ −1.0 × 10 ⁻⁶ NE ³ + 7.0 × 10 ⁻⁴ NE
Then, the relational expression between ε and IVC in the 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 the layer 2 as illustrated in FIG.

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

Therefore, in the case of determining the valve closing timing IVC in the engine 1 of low octane fuel, the designer sets the geometric compression ratio ε to 10 ≦ ε <15.7, and when the engine 1 has a rotational speed of 2000 rpm,
0.4234ε ² −21.826ε + 178.75 ≦ IVC ≦ −0.4234ε ² + 21.826ε−138.75 (6)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 15.7 ≦ ε <18.7, and when the rotational 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 determined so as to satisfy the above.

Further, in the engine 1 of the low octane fuel, the designer determines that the geometric compression ratio ε is 18.7 ≦ ε ≦ 30, and when the rotational speed of the engine 1 is 2000 rpm,
−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 determined so as to satisfy the above.

  The range where the cross-hatching is given to the graph 1702 in FIG. 17 is a range where the ε-IVC establishment range of the high octane number fuel and the ε-IVC establishment range of the low octane number fuel overlap. When the designer determines the IVC within a range where the two formation ranges overlap, the designer can set 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 varies from destination to destination, the designer can design the engine control logic in a batch. Collective design has the advantage of reducing design man-hours.

In addition, although illustration is omitted, even in the engine 1 of low octane fuel, the ε-IVC establishment range moves parallel to the retard side when the rotational speed of the engine 1 increases. In the engine 1 with a low octane fuel, the designer defines the geometric compression ratio ε as 10 ≦ ε <15.7, and when the engine 1 has a rotational speed of 3000 rpm,
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 determined so as to satisfy the above.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 15.7 ≦ ε <18.7, and when the rotational 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 determined so as to satisfy the above.

Further, when the designer determines that the geometric compression ratio ε is 18.7 ≦ ε ≦ 30 in the engine 1 with a low octane fuel, the engine 1 has a rotational speed of 3000 rpm,
−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 determined so as to satisfy the above.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 10 ≦ ε <15.7, and when the engine 1 has a rotational speed of 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 determined so as to satisfy the above.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 15.7 ≦ ε <18.7, and when the rotational 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 determined so as to satisfy the above.

Further, when the designer determines that the geometric compression ratio ε is 18.7 ≦ ε ≦ 30 in the engine 1 of the low octane fuel, when the rotational speed of the engine 1 is 4000 rpm,
−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 determined so as to satisfy the above.

Similarly to the above, when the correction term C related to the rotational speed NE (rpm) of the engine 1 is used, the relational expression between ε and IVC in the layer 2 can be expressed as follows in the engine 1 of the low octane fuel. it 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 geometric compression ratio in layer 3 and closing timing of intake valve)
FIG. 18 shows the characteristics of SPCCI combustion when the engine 1 performs SPCCI combustion of an air-fuel mixture that is leaner than the stoichiometric air-fuel ratio in the same manner as the 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 combustion center of gravity, and the combustion center of gravity advances as it goes upward in the figure.

  The range in which the SPCCI combustion is stable is a range surrounded by a curve in FIG. When the load of the engine 1 is low, a range in which SPCCI combustion is stable is located at the upper left in FIG. When the load on the engine 1 increases, the range in which SPCCI combustion is stabilized moves downward in FIG.

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

  As can be seen from the figure, when the load on the engine 1 is high, the area where the range in which the combustion stability is secured and the range in which the emission of RawNOx can be suppressed overlaps. On the other hand, when the load of the engine 1 is low, the area where the range in which the combustion stability is secured and the range in which the discharge of RawNOx can be suppressed overlaps.

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

The present inventors have, like the above, in actual engine 1, the measuring various parameters during operation at light load in the layer 3, from the measured parameters, the actual combustion chamber 17 in theta _CI The temperature of was estimated. And the average value of the estimated several temperature was defined as reference temperature Tth3. If the reference temperature Tth3 temperature of the combustion chamber 17 in the theta _CI, it is possible to realize a SPCCI combustion in the layer 3. This reference temperature Tth3 corresponds to the lower limit temperature. When the engine 1 is operating at light load, theta if the temperature in the combustion chamber 17 is the reference temperature Tth3 or the _CI, it is possible to realize a SPCCI combustion by retarding the ignition timing. However, theta when the temperature in the combustion chamber 17 in the _CI is less than the reference temperature Tth3, can not be realized SPCCI combustion be advanced ignition timing.

Therefore, in order to achieve a stable SPCCI combustion in the layer 3, the temperature in the combustion chamber 17 in the theta _CI is such that the reference temperature Tth3 above, the relationship between ε and IVC are required.

  Therefore, as conceptually shown in FIG. 19, the inventors of the present application, the valve closing timing IVC of the intake valve 21 for each geometric compression ratio ε (ε1, ε2,...), The intake valve 21 and Calculation of the temperature of the combustion chamber 17 during CI combustion using the model of the 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 numeral 1901). As indicated by hatching in the matrix 1901, the combination of IVC and O / L that is equal to or higher than the reference temperature Tth 3 can realize SPCCI combustion in the 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 indicated by reference numeral 1902 in FIG. 19, the inventors of the present application change values for each of IVC and O / L in a matrix composed of two parameters of valve closing timing IVC and overlap period O / L. However, G / F estimation calculation was performed using the model of the engine 1. As indicated by hatching in the matrix denoted by reference numeral 1902, the combination of IVC and O / L in which G / F is equal to or greater than a predetermined value can suppress the discharge of RawNOx.

  Then, the inventors of the present application show a combination of IVC and O / L that is equal to or higher than the reference temperature Tth3 indicated by reference numeral 1901, and IVC and O / L that indicate G / F that is equal to or higher than a predetermined value indicated by reference numeral 1902. By overlapping the combinations, the relationship between ε and IVC, which can achieve both the stability of SPCCI combustion and the suppression of RawNOx emission, was defined. In other words, in the matrix denoted by reference numeral 1903, the cross hatched range is a combination of ε and IVC that can realize both the stability of SPCCI combustion and the suppression of NOx emission.

  Although not shown in the drawings, the inventors of the present application similarly estimate the temperature of the combustion chamber 17 during CI combustion and G for the matrix in which the 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 Tth3 or higher and a combination of IVC and O / L having a G / F of a predetermined value or higher were obtained.

  FIG. 20 shows approximate expressions (V) and (VI) calculated from a combination of ε and IVC. The horizontal axis in 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 diagram 2001 in FIG. 20 corresponds to the case where the rotational speed of the engine 1 is 2000 rpm. The approximate expressions (V) and (VI) are respectively
Approximate expression (V) IVC = −0.9949ε ² + 41.736ε−361.16
Approximate Formula (VI) IVC = 0.9949ε ² −41.736ε + 401.16
It is.

  In FIG. 20, the combination of ε and IVC on the right side of the approximate expressions (V) and (VI) is such that the temperature of the combustion chamber 17 during CI combustion is equal to or higher than the reference temperature Tth3, and 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 that can realize SPCCI combustion when the engine 1 is operated at a light load in the layer 3.

  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, the combustion concept of SPCCI combustion is that, as described above, 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 burns CI. . From the examination of the experiments and the like conducted by the present inventors so far, the self-ignition of the air-fuel mixture occurs when the ambient temperature of the air-fuel mixture that self-ignites exceeds a predetermined reference temperature Tth4. The reference temperature Tth4 at this time is It has been found that the average temperature of the entire combustion chamber 17 does not necessarily match. From this knowledge, when the average temperature in the combustion chamber 17 at the compression top dead center reaches the reference temperature Tth4, it is considered that the CI combustion starts before the SI combustion starts. In this case, the SPCCI combustion is performed. I can't do it.

  Accordingly, the inventors of the present application, like the matrix 1901 in FIG. 19, the valve closing timing IVC of the intake valve 21 for each geometric compression ratio ε (ε1, ε2,...) And the intake valve 21. In the matrix of the overlap period O / L of the exhaust valve 22 and the valve closing timing IVC and the overlap period O / L, the value of the valve closing timing IVC and the overlap period O / L is changed. The temperature of the combustion chamber 17 was estimated and calculated. The combination of IVC and O / L in which the temperature in the combustion chamber 17 at the time of compression top dead center exceeds the reference temperature Tth4 cannot realize SPCCI combustion, and IVC and O / L below the reference temperature Tth4. This combination can achieve SPCCI combustion.

FIG. 20 shows approximate expressions (VII) and (VIII) calculated from a combination of ε and IVC at which the temperature in the combustion chamber 17 at the compression top dead center is equal to or lower than the reference temperature Tth4. The approximate expressions (VII) and (VIII) are respectively
Approximate Formula (VII) IVC = −4.7481ε ² + 266.75ε−3671.2
And approximate formula (VIII) IVC = 4.77481ε ² −266.75ε + 3711.2
It is.

  In FIG. 20, the combination of ε and IVC on the left side of the approximate expressions (VII) and (VIII) can avoid the start of CI combustion 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 symmetric in the vertical direction with IVC = about 20 deg. Further, IVC = 75 deg. ABDC is a delay limit of the 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 the layer 3, and IVC = Similarly, −35 deg. ABDC is an advance limit of the 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 operates in the layer 3, ε surrounded by the approximate expressions (V), (VI), (VII), and (VIII) in FIG. -IVC must be determined within the IVC formation range (the hatched range in FIG. 20).

Specifically, the designer sets the geometric compression ratio ε to 10 ≦ ε <20, and when the rotational 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 determined so as to satisfy the above.

The designer also sets the geometric compression ratio ε to 20 ≦ ε <25, and when the engine 1 has a rotational speed of 2000 rpm,
−35 ≦ IVC ≦ 75 (12)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Furthermore, when the designer defines the geometric compression ratio ε as 25 ≦ ε ≦ 30, the engine 1 has a rotational speed of 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 determined so as to satisfy the above.

  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 air-fuel mixture whose A / F is leaner than the stoichiometric air-fuel ratio is subjected to SPCCI combustion. The valve closing timing IVC is set for each operation state determined by the load and rotation speed of the engine 1 in the layer 3.

The example shown by the solid line in FIG. 20 is the ε-IVC establishment range when the rotational speed of the engine 1 is 2000 rpm as described above. When the rotation speed of the engine 1 changes, the ε-IVC establishment range also changes. The point that the ε-IVC establishment range moves parallel to the retard side when the rotational speed of the engine 1 increases is the same in FIG. Therefore, the designer sets the geometric compression ratio ε to 10 ≦ ε <20, and when the rotational speed of the engine 1 is 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 determined so as to satisfy the above.

The designer also sets the geometric compression ratio ε to 20 ≦ ε <25, and when the engine 1 has a rotational speed of 3000 rpm,
−33 ≦ IVC ≦ 77 (12 ⁻¹ )
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Furthermore, if the designer determines the geometric compression ratio ε to be 25 ≦ ε ≦ 30, the engine 1 has a rotational speed of 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 determined so as to satisfy the above.

The designer also defines the geometric compression ratio ε as 10 ≦ ε <20, and when the rotation speed of the engine 1 is 4000 rpm (see the one-dot chain 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 determined so as to satisfy the above.

The designer also sets the geometric compression ratio ε to 20 ≦ ε <25, and when the rotation speed of the engine 1 is 4000 rpm,
−27 ≦ IVC ≦ 83 (12 ⁻² )
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Furthermore, when the designer sets the geometric compression ratio ε to 25 ≦ ε ≦ 30, the engine 1 has a rotational speed of 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 determined so as to satisfy the above.

Similarly to the above, when the correction term C relating to the rotational speed NE (rpm) of the engine 1 is used, the relational expression between ε and IVC in the 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 the layer 3 as illustrated in FIG.

  A lower diagram 2002 in FIG. 20 shows a relationship between ε and IVC when the fuel is a low-octane fuel.

In the engine 1 with a low octane fuel, the designer determines the valve closing timing IVC. If the geometric compression ratio ε is set to 10 ≦ ε <18.7, the engine 1 has a rotational speed of 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 determined so as to satisfy the above.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 18.7 ≦ ε <23.7, and when the engine 1 has a rotational speed of 2000 rpm,
−35 ≦ IVC ≦ 75 (16)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

Further, the designer determines that the geometric compression ratio ε is 23.7 ≦ ε ≦ 30 in the engine 1 of low octane fuel, and when the rotational speed of the engine 1 is 2000 rpm,
−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 determined so as to satisfy the above.

  The range indicated by the cross hatching in the lower diagram 2002 of FIG. 20 is a range where the ε-IVC establishment range of the high octane fuel and the ε-IVC establishment range of the low octane fuel overlap. Similarly to the above, when the IVC is determined within a range where the two establishment ranges overlap, the designer sets 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.

In addition, although illustration is omitted, even in the engine 1 of low octane fuel, the ε-IVC establishment range moves parallel to the retard side when the rotational speed of the engine 1 increases. In the engine 1 with a low octane fuel, the designer defines the geometric compression ratio ε as 10 ≦ ε <18.7, and when the engine 1 has a rotational speed of 3000 rpm,
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 determined so as to satisfy the above.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 18.7 ≦ ε <23.7, and when the engine 1 has a rotational speed of 3000 rpm,
−33 ≦ IVC ≦ 77 (16 ⁻¹ )
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 23.7 ≦ ε ≦ 30, and when the rotational 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 determined so as to satisfy the above.

In the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 10 ≦ ε <18.7, and when the engine 1 has a rotational speed of 4000 rpm,
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 determined so as to satisfy the above.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 18.7 ≦ ε <23.7, and when the rotational speed of the engine 1 is 4000 rpm,
−27 ≦ IVC ≦ 83 (16 ⁻² )
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 23.7 ≦ ε ≦ 30, and when the rotational speed of the engine 1 is 4000 rpm,
−3.1298ε ² + 172.48ε−2292 ≦ IVC ≦ 83 (17 ⁻² )
Or
−27 ≦ IVC ≦ 3.1298ε ² −172.48ε + 2348 (18 ⁻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 rotational speed NE (rpm) of the engine 1 is used, the relational expression between ε and IVC in the layer 3 can be expressed as follows in the engine 1 of the low octane fuel. it 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 geometric compression ratio in layer 2 and layer 3 and closing timing of intake valve)
FIG. 21 shows a relationship between the geometric compression ratio ε and the valve closing timing IVC of the intake valve 21 that enables SPCCI combustion in both the layer 2 and the 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 operation 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 the layer 2 and the layer 3, even when the map of the engine 1 is switched from the layer 2 to the layer 3, the SPCCI It becomes possible to continue the combustion.

  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 fuel.

In the engine 1 with a high octane fuel, the designer determines that the geometric compression ratio ε is 10 ≦ ε <17, and when the rotational speed of the engine 1 is 2000 rpm,
0.9949ε ² −41.736ε + 401.16 ≦ IVC ≦ −0.9949ε ² + 41.736ε−361.16 (19)
The valve closing timing IVC (deg. ABDC) is determined so as to satisfy the above.

In the engine 1 of high octane fuel, the designer determines that the geometric compression ratio ε is 17 ≦ ε ≦ 30, and when the rotational speed of the engine 1 is 2000 rpm,
−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 determined so as to satisfy the above.

  By setting the valve closing timing IVC of the intake valve 21 based on the relational expressions (19) to (21), the air-fuel mixture in which the A / F is leaner than the stoichiometric air-fuel ratio can be subjected to SPCCI combustion, SPCCI combustion can be performed on an air / fuel mixture in which A / F is richer than the stoichiometric air / fuel ratio or G / F is leaner than the stoichiometric air / fuel ratio.

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

As shown by the broken line, in the engine 1 of high octane fuel, the designer determines that the geometric compression ratio ε is 10 ≦ ε <17, and when the rotational speed of the engine 1 is 3000 rpm,
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 determined so as to satisfy the above.

In the engine 1 of high octane fuel, the designer determines that the geometric compression ratio ε is 17 ≦ ε ≦ 30. When the engine 1 has a rotational speed of 3000 rpm,
−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 determined so as to satisfy the above.

Similarly, as indicated by the alternate long and short dash line, in the engine 1 of high octane fuel, the designer also defines the geometric compression ratio ε as 10 ≦ ε <17, and when the rotational 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 determined so as to satisfy the above.

In the engine 1 of high octane fuel, the designer determines that the geometric compression ratio ε is 17 ≦ ε ≦ 30. When the engine 1 has a rotational speed of 4000 rpm,
−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 determined so as to satisfy the above.

Similarly to the above, when the correction term C related to the rotational speed NE (rpm) of the engine 1 is used, the relational expression between ε and IVC in the layer 2 and the 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 determined to be 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 increases.

Further, as shown in the lower diagram 2102 of FIG. 21, in the engine 1 of the low octane fuel, the designer determines that the geometric compression ratio ε is 10 ≦ ε <15.7, and the rotational 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.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 15.7 ≦ ε ≦ 30, and when the rotational speed of the engine 1 is 2000 rpm,
−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 illustration is omitted, in the engine 1 with a low octane number fuel, the designer defines the geometric compression ratio ε as 10 ≦ ε <15.7, and when the rotational 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.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 15.7 ≦ ε ≦ 30, and when the rotational speed of the engine 1 is 3000 rpm,
−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, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 10 ≦ ε <15.7,
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.

In addition, in the engine 1 with a low octane fuel, the designer determines that the geometric compression ratio ε is 15.7 ≦ ε ≦ 30, and when the rotational speed of the engine 1 is 4000 rpm,
−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 rotational speed NE (rpm) of the engine 1 is used, in the low octane fuel engine 1, the relational expression between ε and IVC in the layers 2 and 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 illustration is omitted, the designer may determine the IVC within a range where the ε-IVC establishment range in the upper diagram 2101 in FIG. 21 overlaps the ε-IVC establishment range in the lower diagram 2102. Similarly to the above, when the IVC is determined within a range where the two establishment ranges overlap, the designer sets 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 for designing the control logic of the engine 1 that performs SPCCI combustion will be described with reference to the flowchart shown in FIG. The designer can execute each step using a computer. The computer stores information on the ε-IVC establishment range illustrated in FIGS. 17, 20, and 21.

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

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

  In step S223, the designer sets an operating state composed of the load and the rotational speed of the engine 1, and in subsequent step S224, the designer selects the ε-IVC establishment range stored in the computer (FIGS. 17, 20, and 21). ) To select IVC.

  In step S225, the computer determines whether SPCCI combustion can be realized based on the IVC set in step S224. If the determination in step S225 is yes, the process moves to step S226, and the designer defines the control logic of the engine 1 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 to execute SI combustion in the operating state set in step S223. In step S227, the designer may newly set the closing timing IVC of the intake valve 21 in consideration of performing SI combustion.

  As described above, the method for designing the control logic of the compression ignition engine disclosed herein 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 satisfying the relationship. The designer can design the control logic of the engine 1 with fewer man-hours than in the past.

(Other embodiments)
The technique disclosed here 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 include a turbocharger instead of the mechanical supercharger 44.

1 Engine 10 ECU (control unit)
17 Combustion chamber 23 Intake motorized 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 First intake air temperature sensor (measurement unit)
SW3 First pressure sensor (measurement unit)
SW4 Second intake air temperature sensor (measurement unit)
SW5 Second pressure sensor (measurement unit)
SW6 Acupressure 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 Third intake air temperature sensor (measurement unit)

Claims (10)

A method of designing the control logic of a compression ignition engine,
The engine is
A fuel injection section for injecting fuel to be supplied into the combustion chamber;
A variable valve mechanism that changes the valve timing of the intake valve;
An ignition part for igniting the air-fuel mixture in the combustion chamber;
A measuring unit for measuring a parameter related to the operating state of the engine;
A control unit that receives a measurement result of the measurement unit and performs calculation according to a control logic corresponding to an operating state of the engine, and outputs a signal to the fuel injection unit, the variable valve mechanism, and the ignition unit; With
The engine is
The control unit sends a signal 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 in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. And a first mode for outputting a signal to the ignition unit so that the unburned mixture burns by self-ignition after the ignition unit ignites the mixture in the combustion chamber;
The control unit determines that G / F, which is a weight ratio of the total gas in the mixture in the combustion chamber and the fuel, is leaner than the stoichiometric air-fuel ratio, and the A / F is richer than the stoichiometric air-fuel ratio or stoichiometric air-fuel ratio. So that the unburned mixture is burned by self-ignition after the ignition unit has ignited the mixture in the combustion chamber. A second mode for outputting a signal to the unit,
The method of designing the control logic is as follows:
Determining a geometric compression ratio ε of the engine;
Determining a control logic that defines the closing timing IVC of the intake valve,
When setting the control logic, 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
Where C is a correction term relating to the engine speed NE (rpm),
C = 3.3 × 10 ⁻¹⁰ NE ³ −1.0 × 10 ⁻⁶ NE ³ + 7.0 × 10 ⁻⁴ NE
The control logic of the compression ignition engine that defines the valve closing timing IVC (deg. ABDC) is designed so as to satisfy the following equation. A method of designing the control logic of a compression ignition engine,
The engine is
A fuel injection section for injecting fuel to be supplied into the combustion chamber;
A variable valve mechanism that changes the valve timing of the intake valve;
An ignition part for igniting the air-fuel mixture in the combustion chamber;
A measuring unit for measuring a parameter related to the operating state of the engine;
A control unit that receives a measurement result of the measurement unit and performs calculation according to a control logic corresponding to an operating state of the engine, and outputs a signal to the fuel injection unit, the variable valve mechanism, and the ignition unit; With
The engine is
The control unit sends a signal 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 in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. And a first mode for outputting a signal to the ignition unit so that the unburned mixture burns by self-ignition after the ignition unit ignites the mixture in the combustion chamber;
The control unit determines that G / F, which is a weight ratio of the total gas in the mixture in the combustion chamber and the fuel, is leaner than the stoichiometric air-fuel ratio, and the A / F is richer than the stoichiometric air-fuel ratio or stoichiometric air-fuel ratio. So that the unburned mixture is burned by self-ignition after the ignition unit has ignited the mixture in the combustion chamber. A second mode for outputting a signal to the unit,
The method of designing the control logic is as follows:
Determining a geometric compression ratio ε of the engine;
Determining a control logic that defines the closing timing IVC of the intake valve,
When setting the control logic, 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
Or
0.9949ε ² −41.736ε + 401.16 + C ≦ IVC ≦ 1.9163ε ² −89.935ε + 974.94 + C
Where C is a correction term relating to the engine speed NE (rpm),
C = 3.3 × 10 ⁻¹⁰ NE ³ −1.0 × 10 ⁻⁶ NE ³ + 7.0 × 10 ⁻⁴ NE
The control logic of the compression ignition engine that defines the valve closing timing IVC (deg. ABDC) is designed so as to satisfy the following equation. A method of designing the control logic of a compression ignition engine,
The engine is
A fuel injection section for injecting fuel to be supplied into the combustion chamber;
A variable valve mechanism that changes the valve timing of the intake valve;
An ignition part for igniting the air-fuel mixture in the combustion chamber;
A measuring unit for measuring a parameter related to the operating state of the engine;
A control unit that receives a measurement result of the measurement unit and performs calculation according to a control logic corresponding to an operating state of the engine, and outputs a signal to the fuel injection unit, the variable valve mechanism, and the ignition unit; With
The engine is
The control unit sends a signal 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 in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. And a first mode for outputting a signal to the ignition unit so that the unburned mixture burns by self-ignition after the ignition unit ignites the mixture in the combustion chamber;
The control unit determines that G / F, which is a weight ratio of the total gas in the mixture in the combustion chamber and the fuel, is leaner than the stoichiometric air-fuel ratio, and the A / F is richer than the stoichiometric air-fuel ratio or stoichiometric air-fuel ratio. So that the unburned mixture is burned by self-ignition after the ignition unit has ignited the mixture in the combustion chamber. A second mode for outputting a signal to the unit,
The method of designing the control logic is as follows:
Determining a geometric compression ratio ε of the engine;
Determining a control logic that defines the closing timing IVC of the intake valve,
When setting the control logic, if the fuel is a low octane fuel and 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
Where C is a correction term relating to the engine speed NE (rpm),
C = 3.3 × 10 ⁻¹⁰ NE ³ −1.0 × 10 ⁻⁶ NE ³ + 7.0 × 10 ⁻⁴ NE
The control logic of the compression ignition engine that defines the valve closing timing IVC (deg. ABDC) is designed so as to satisfy the following equation. A method of designing the control logic of a compression ignition engine,
The engine is
A fuel injection section for injecting fuel to be supplied into the combustion chamber;
A variable valve mechanism that changes the valve timing of the intake valve;
An ignition part for igniting the air-fuel mixture in the combustion chamber;
A measuring unit for measuring a parameter related to the operating state of the engine;
A control unit that receives a measurement result of the measurement unit and performs calculation according to a control logic corresponding to an operating state of the engine, and outputs a signal to the fuel injection unit, the variable valve mechanism, and the ignition unit; With
The engine is
The control unit sends a signal 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 in the combustion chamber to the fuel, is leaner than the stoichiometric air-fuel ratio. And a first mode for outputting a signal to the ignition unit so that the unburned mixture burns by self-ignition after the ignition unit ignites the mixture in the combustion chamber;
The control unit determines that G / F, which is a weight ratio of the total gas in the mixture in the combustion chamber and the fuel, is leaner than the stoichiometric air-fuel ratio, and the A / F is richer than the stoichiometric air-fuel ratio or stoichiometric air-fuel ratio. So that the unburned mixture is burned by self-ignition after the ignition unit ignites the mixture in the combustion chamber. A second mode for outputting a signal to the unit,
The method of designing the control logic is as follows:
Determining a geometric compression ratio ε of the engine;
Determining a control logic that defines the closing timing IVC of the intake valve,
When setting the control logic, if the fuel is a low octane fuel and 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
Or
0.9949ε ² −39.149ε + 348.59 + C ≦ IVC ≦ 1.9211ε ² −85.076ε + 862.01 + C
Where C is a correction term relating to the engine speed NE (rpm),
C = 3.3 × 10 ⁻¹⁰ NE ³ −1.0 × 10 ⁻⁶ NE ³ + 7.0 × 10 ⁻⁴ NE
The control logic of the compression ignition engine that defines the valve closing timing IVC (deg. ABDC) is designed so as to satisfy the following equation. In the method of designing the control logic of the compression ignition type engine according to any one of claims 1 to 4,
The closing timing IVC of the intake valve changes when the engine operating state changes,
A method of designing a control logic of a compression ignition engine that determines the valve closing timing IVC (deg. ABDC) so as to satisfy the relational expression for each operating state. In the method of designing the control logic of the compression ignition engine according to any one of claims 1 to 5,
The engine is a method for designing a control logic of a compression ignition engine that operates in a low load operation state of a predetermined load or less according to the control logic. A method for designing the control logic of a compression ignition engine according to claim 6,
The engine is a method for designing a control logic of a compression ignition engine that operates in a minimum load operation state according to the control logic. A method for designing a control logic of a compression ignition engine according to any one of claims 1 to 7,
The controller is in a first mode when the wall temperature of the combustion chamber is equal to or higher than a first predetermined temperature and the temperature of the intake air is equal to or higher than a second predetermined temperature, and the wall temperature of the combustion chamber is set to the first predetermined temperature. When the temperature of the intake air is less than the second predetermined temperature, the control logic of the compression ignition engine that is set to the second mode is designed. A method for designing control logic of a compression ignition engine according to any one of claims 1 to 8,
The engine includes an EGR system that introduces exhaust gas into the combustion chamber,
The control unit has a heat quantity ratio as an index related to a ratio of heat quantity generated when the air-fuel mixture is combusted by flame propagation with respect to the total heat quantity generated when the air-fuel mixture in the combustion chamber burns, A method of designing a control logic of a compression ignition engine that outputs a signal to the EGR system and the ignition unit so as to achieve a target heat amount ratio determined in accordance with an operation state of the engine. A method for designing the control logic of a compression ignition engine according to claim 9,
The control unit designs a control logic of a compression ignition engine that outputs a signal to the EGR system and the ignition unit so that the heat ratio is higher when the engine load is high than when the engine load is low.

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