JP2018193987A - Control device for compression ignition type engine - Google Patents

Abstract

To suppress the NVH of a compression ignition type engine to a permissible value or lower during high revolution.SOLUTION: A control device for the compression ignition type engine includes a spark plug 25, a controller 10, and an EGR system 55 for introducing burnt gas from an exhaust passage 50 via an intake passage 40 into a combustion chamber 17. The controller 10 outputs a control signal to the spark plug 25 at a predetermined timing so that self-ignition combustion of unburnt air-fuel mixture is caused by firing. The controller 10 outputs a control signal to the EGR system 55 so that an EGR rate as the rate of the amount of the burnt gas to that of air-fuel mixture in the combustion chamber 17 becomes lower when the revolution speed of an engine 1 is high than when low.SELECTED DRAWING: Figure 13

Description

  The technology disclosed herein relates to a control device for a compression ignition engine.

  Patent Document 1 describes an engine that burns an air-fuel mixture in a combustion chamber by self-ignition in a predetermined region of low load and low rotation. The engine burns the air-fuel mixture by spark ignition in a region where the load is higher than the region where the low load is low and the region where the rotational speed is higher than the region where the low load is low.

Japanese Patent No. 4082292

  By the way, combustion by compression ignition generates a relatively large combustion noise. When the engine speed is high, the NVH (Noise Vibration Harshness) of the engine exceeds an allowable value.

  The technology disclosed herein has been made in view of such a point, and an object thereof is to perform combustion by compression ignition while suppressing NVH below an allowable value in a compression ignition engine.

  As described later, the inventors of the present application have considered a combustion mode (SPCCI combustion) that combines SI (Spark Ignition) combustion and CI (Compression Ignition) combustion. That is, the air-fuel mixture in the combustion chamber is forcibly ignited and burned by flame propagation, and the unburned air-fuel mixture in the combustion chamber is combusted by self-ignition due to the heat generated by SI combustion.

  In combustion by self-ignition, if the temperature in the combustion chamber before the start of compression varies, the self-ignition timing changes greatly. For example, if the timing of self-ignition is advanced, there is a problem that combustion noise increases.

  In SPCCI combustion, variation in temperature in the combustion chamber before the start of compression can be absorbed by changing the calorific value of SI combustion. If the start timing of SI combustion is changed by changing the ignition timing, for example, according to the temperature in the combustion chamber before the compression starts, the self-ignition timing can be controlled.

  Combustion by flame propagation can suppress the generation of combustion noise because the pressure fluctuation is relatively small. Also, by performing CI combustion, the combustion period is shortened and combustion efficiency is improved compared to combustion by flame propagation. Therefore, the combustion mode combining SI combustion and CI combustion can improve fuel efficiency while suppressing the generation of combustion noise.

  If SPCCI combustion is performed when the engine speed is high, CI combustion can be performed while NVH is kept below an allowable value.

  Specifically, the technology disclosed herein is a control apparatus for a compression ignition engine, and is arranged to face an engine configured to burn an air-fuel mixture in a combustion chamber. And an ignition plug configured to ignite an air-fuel mixture in the combustion chamber, and an exhaust passage attached to the engine and exhausting burned gas generated in the combustion chamber to the combustion chamber. Each of the EGR system configured to introduce the burned gas into the combustion chamber by way of an intake passage through which gas flows in, the spark plug, and the EGR system; and A spark plug and a controller configured to operate the engine by outputting a control signal to each of the EGR systems.

  The controller outputs a control signal to the spark plug at a predetermined ignition timing so that a predetermined form of combustion is performed in which the unburned mixture in the combustion chamber is combusted by self-ignition by ignition of the spark plug. . The EGR system is configured such that when the engine speed is higher than when the engine speed is low, an EGR rate that is a ratio of the amount of the burned gas contained in the air-fuel mixture in the combustion chamber is reduced. Output a control signal.

  The “combustion chamber” here is not limited to the meaning of the space when the piston reaches compression top dead center. The term “combustion chamber” is used in a broad sense.

  According to this configuration, when the controller determines that it is necessary from the operating state of the engine, it outputs a control signal to the spark plug. Thereby, the spark plug ignites the air-fuel mixture formed in the combustion chamber with a predetermined ignition wing. Thereby, combustion is started, and the unburned mixture is combusted by self-ignition following the combustion. That is, the aforementioned SPCCI combustion is performed.

  When the SPCCI combustion is performed, the controller outputs a control signal to the EGR system attached to the engine. This EGR system passes burned gas into the combustion chamber by passing from the exhaust passage (passage for discharging burned gas generated in the combustion chamber) through the intake passage (passage for allowing unburned gas to flow into the combustion chamber). That is, it is configured to introduce EGR gas (so-called external EGR). Therefore, since this EGR gas is indirectly introduced into the combustion chamber via a passage outside the engine, its temperature becomes relatively low. It is also possible to forcibly cool. That is, the temperature of the EGR gas introduced by the EGR system can be lower than that in the combustion chamber.

  The EGR system is configured to change the EGR rate (the ratio of the amount of EGR gas contained in the air-fuel mixture in the combustion chamber), and the engine speed is determined according to the control signal output from the controller. The EGR rate is made lower when it is higher than when it is low.

  As described above, NVH becomes larger at the time of high engine rotation than at low speed. Therefore, as in the case of low rotation at high rotation, when CI combustion with large combustion noise is performed, NVH may exceed an allowable value. Therefore, in order to perform combustion by self-ignition at high speed, it is necessary to reduce CI combustion in SPCCI combustion and increase SI combustion.

  Further, the combustion cycle is shortened at the time of high rotation compared to that at the time of low rotation. Accordingly, the amount of heat received by the inner wall of the combustion chamber during combustion is reduced and the temperature is lowered, so that the air-fuel mixture is less likely to receive heat and the time for which the air-fuel mixture burns is also shortened. Therefore, at the time of high rotation, under such unfavorable conditions, it is necessary to promote SPCCI combustion and increase the rate of SI combustion. However, if the temperature of the air-fuel mixture itself (temperature immediately before combustion) is increased, SPCCI combustion can be promoted even under such conditions.

  In the SPCCI combustion, the CI combustion is ignited after the SI combustion. When the temperature of the air-fuel mixture is increased, both SI combustion and CI combustion are promoted. However, since the combustion time is short at a high rotation speed, the SI combustion in which combustion starts first becomes faster and the CI in which combustion starts later. Combustion is promoted rather than combustion. That is, by increasing the temperature of the air-fuel mixture, it is possible to increase the rate of SI combustion at the time of high rotation.

  Therefore, when the engine is in operation, the EGR rate is lowered and the ratio of EGR gas (low-temperature burned gas) contained in the mixture itself can be reduced, so that the temperature of the mixture can be increased. In combustion, SI combustion can be increased. As a result, NVH can be suppressed to an allowable value or less even during high rotation.

  An intake air flow control device attached to the engine and configured to change the flow of gas introduced into the combustion chamber, the controller when the engine speed is higher than when it is low The control signal may be output to the EGR system so that the EGR rate is low, and the control signal may be output to the intake air flow control device so that the gas flow is strong.

  The EGR rate is limited by other state quantities such as G / F of the air-fuel mixture. Therefore, only by changing the EGR rate, the suppression effect of combustion noise in SPCCI combustion is limited, and a sufficient effect may not be obtained. On the other hand, if the intake air flow is strengthened, vaporization of the injected fuel is promoted, and SI combustion can be performed in a state where the flow in the combustion chamber is strong. Thereby, SI combustion becomes rapid, and SI combustion is promoted as compared with CI combustion that ignites with a delay. Therefore, by lowering the EGR rate and strengthening the intake air flow, it is possible to compensate for the effects that are limited by changing the EGR rate. As a result, the ratio of SI combustion in SPCCI combustion can be increased, and NVH can be suppressed to an allowable value or less even at high revolutions.

  The controller outputs a control signal to the EGR system so that the EGR rate becomes lower when the engine speed is higher than when the engine speed is low, and sends a control signal to the spark plug so that the ignition timing is advanced. May be output.

  If the ignition timing is advanced, the start of SI combustion is accelerated. Therefore, since SI combustion becomes rapid even in a short time at high rotation, the rate of SI combustion in SPCCI combustion can be increased. Therefore, the ignition timing advance angle can also compensate for the effect that is limited by the change in the external EGR rate, and NVH can be suppressed to an allowable value or less even at high revolutions.

  The controller may output a control signal to the EGR system so that the EGR rate starts to decrease when the engine speed exceeds a preset reduction start speed.

  By doing so, the EGR rate can be set to a high value at a low rotation speed that is equal to or less than the lowering start rotation speed. / F can be kept within an appropriate range. Since the dilution ratio of the air-fuel mixture can be kept high, the fuel efficiency can be improved. And at the time of the high rotation exceeding a fall start rotation speed, the ratio of the amount of EGR gas introduced into the combustion chamber decreases, and the temperature of the air-fuel mixture increases. As a result, SI combustion becomes rapid and CI combustion decreases, so that combustion noise generated by SPCCI combustion can be suppressed. NVH can be suppressed to an allowable value or less during high rotation.

  A supercharging system attached to the engine and configured to supercharge a gas introduced into the combustion chamber; and an internal EGR gas provided in the engine and included in an air-fuel mixture in the combustion chamber An internal EGR system configured to change an internal EGR rate that is a percentage of the amount of the engine, wherein the controller is in a first region with a higher load than a predetermined load. A control signal is output to the supercharging system so that the supercharging system does not perform supercharging when in the second region below the predetermined load, and the controller is at least in the first region, A control signal is output to the EGR system so that the EGR rate decreases when the engine speed is higher than when the engine speed is low, and in the second region, the control signal is higher than when the engine speed is low. So that the internal EGR rate at the time of rolling is increased, and outputs a control signal to the internal EGR system may be.

  The temperature of the air-fuel mixture can also be increased by increasing the internal EGR rate and increasing the proportion of internal EGR gas (high-temperature burned gas) contained in the air-fuel mixture, so that SI combustion in SPCCI combustion can be increased.

  However, in the case of such an engine, in the first region where supercharging is performed, even if the internal EGR gas is introduced, scavenging is performed by the supercharging pressure. Therefore, it is difficult to change the internal EGR rate. Therefore, the controller outputs a control signal so that the internal EGR rate becomes high in the second region where supercharging is not performed. On the other hand, since the EGR gas is introduced through the intake passage, the EGR rate can be changed relatively easily even when supercharging is performed. Accordingly, the controller outputs a control signal so that the EGR rate becomes low in the first region where supercharging is performed. As a result, NVH can be suppressed to an allowable value or less even at high revolutions, and the fuel efficiency of the engine can be improved.

  The spark plug is disposed so as to face an upper central portion of the combustion chamber, is provided in the engine, and is an internal EGR rate that is a ratio of the amount of internal EGR gas contained in the air-fuel mixture in the combustion chamber An internal EGR system configured to change the flow rate, and when the predetermined form of combustion is performed, a swirl flow is formed in the combustion chamber, and the controller has a low engine speed. It is good also as outputting a control signal to the internal EGR system so that the internal EGR rate may become high when it is higher.

  When a swirl flow is formed in the combustion chamber in which the spark plug is disposed so as to face the upper central portion of the combustion chamber, the flow of the air-fuel mixture in the combustion chamber is promoted, and the stratification in the combustion chamber A gasified mixture can be formed. As a result, even if the air-fuel mixture in the combustion chamber contains a large amount of internal EGR gas, the amount of internal EGR gas around the spark plug can be reduced at the ignition timing, and SI combustion can be performed stably. Therefore, stable SPCCI combustion can be realized even if the internal EGR rate becomes high at high revolutions.

  The internal EGR gas may be introduced into the combustion chamber by providing a so-called positive overlap period in which the valve opening period of the intake valve and the valve opening period of the exhaust valve overlap.

  If internal EGR gas is introduced by setting a positive overlap period, pump loss and the like can be reduced.

  The controller outputs a control signal to the EGR system so as to start a decrease in the EGR rate when the engine speed exceeds a preset decrease start speed, and the load of the engine is The change rate of the EGR rate may be kept substantially constant with respect to the change.

  As described above, if the EGR rate is maintained high at the same time as during the low rotation, the temperature of the air-fuel mixture becomes too low, and it may be difficult to achieve both suppression of combustion noise and stable SPCCI combustion. is there. Such a problem can be prevented by reducing the EGR rate when the engine speed exceeds a preset reduction start speed.

  However, in such a case, when the engine load becomes high, the amount of burned gas becomes excessive with respect to the fuel, and the ignition of SI combustion in SPCCI combustion may become unstable. In order to prevent this, it is necessary to increase the decrease rate of the EGR rate as the engine load increases.

  On the other hand, when a swirl flow is formed in the combustion chamber, the fluidity of the air-fuel mixture increases in the combustion chamber. Thereby, the amount of burned gas (EGR gas) around the spark plug where ignition is performed can be reduced. As a result, an excessive increase in burnt gas with respect to the fuel is suppressed, and SPCCI combustion can be performed stably, so that the change rate of the engine load can be reduced without increasing the reduction rate of the EGR rate. Can be held constant.

  According to the disclosed control device for a compression ignition engine, NVH can be suppressed to an allowable value or less even at high revolutions, so that the fuel efficiency of the engine can be improved.

FIG. 1 is a diagram illustrating a configuration of an engine. FIG. 2 is a diagram illustrating the configuration of the combustion chamber, in which the upper diagram is a plan view equivalent view of the combustion chamber, and the lower portion is a II-II cross-sectional view. 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 rig testing apparatus for swirl ratio measurement. FIG. 6 is a diagram illustrating the relationship between the opening ratio of the secondary passage and the swirl ratio. FIG. 7A is a diagram illustrating a first operation region map. FIG. 7B is a diagram illustrating a second operation region map. FIG. 7C is a diagram illustrating a third operation region map. FIG. 8 is a diagram conceptually showing a change in the heat generation rate of SPCCI combustion combining SI combustion and CI combustion. FIG. 9 shows a change in SI rate, a change in state quantity in the combustion chamber, a change in the overlap period of the intake valve and the exhaust valve, and fuel in response to the engine load corresponding to the first operating region map. It is a figure explaining the change of injection timing and ignition timing. The upper diagram of FIG. 10 is a diagram illustrating a change in the combustion waveform with respect to an increase in engine load in non-supercharged SPCCI combustion, and the lower diagram of FIG. 10 is an increase in engine load in supercharged SPCCI combustion. It is a figure which illustrates the change of the combustion waveform with respect to doing. FIG. 11 is a diagram illustrating an example of a relationship between the engine speed and the SI rate in an operation region in which SPCCI combustion of the engine corresponding to the first operation region map is performed. FIG. 12 is a diagram showing an example of the relationship between the engine speed and the internal EGR rate in the operation region in which the SPCCI combustion of the engine corresponding to the first operation region map is performed. (A) is a relationship diagram in the case where the internal EGR gas is introduced by setting the negative overlap period, and (b) is a relationship diagram in the case where the internal EGR gas is introduced by setting the positive overlap period. is there. FIG. 13 is a diagram showing an example of the relationship between the engine speed and the external EGR rate in the operation region in which the SPCCI combustion of the engine corresponding to the first operation region map is performed. (A) is a relationship diagram when a swirl flow is hardly formed in the combustion chamber, and (b) is a relationship diagram when a swirl flow is formed with a predetermined strength in the combustion chamber. It is. FIG. 14 is a diagram illustrating an example of the relationship between the engine speed and the opening of the swirl control valve in the operation region in which the SPCCI combustion of the engine corresponding to the first operation region map is performed. (A) is a relationship diagram in the engine corresponding to the first operation region map, and (b) is a relationship diagram in the engine corresponding to the second or third operation region map. FIG. 15 is a diagram illustrating an example of a relationship between the engine speed and the ignition timing in an operation region in which SPCCI combustion of the engine corresponding to the first operation region map is performed. FIG. 16 is a flowchart showing a procedure of engine control executed by the ECU. FIG. 17 is a diagram illustrating a control concept related to the change of the SI rate. FIG. 18 is a diagram illustrating fuel injection timing, ignition timing, and combustion waveform in each operating state in the third operating region map. FIG. 19 is a simplified diagram corresponding to the third operation region map shown in FIG. 7C. 20 is a diagram illustrating combustion waveforms in the respective operation states shown in FIG.

  Hereinafter, an embodiment of a control device for a compression ignition engine (hereinafter also simply referred to as an engine 1) will be described in detail with reference to the drawings. The following description is an example of the control device for the engine 1. FIG. 1 is a diagram illustrating the configuration of the engine 1. FIG. 2 is a cross-sectional view illustrating the configuration of the combustion chamber. In FIG. 1, the intake side is the left side of the drawing, and the exhaust side is the right side of the drawing. The intake side in FIG. 2 is the right side of the drawing, and the exhaust side is the left side of the drawing. FIG. 3 is a block diagram illustrating the configuration of the control device of the engine 1.

  Further, “EGR gas” used in the following description includes “burned gas (exhaust gas) remaining in the combustion chamber and / or re-inhaled into the combustion chamber”. Similarly, “internal EGR gas” means “burned gas (exhaust gas) that remains in the combustion chamber and / or is directly re-inhaled into the combustion chamber without passing through a passage outside the engine”. The “external EGR gas” includes “burned gas (exhaust gas) indirectly re-inhaled into the combustion chamber via a passage outside the engine such as an exhaust passage or an intake passage”. The “EGR rate” corresponds to “a ratio of the amount of EGR gas contained in the air-fuel mixture (total gas) in the combustion chamber”. The “internal EGR rate” corresponds to “the ratio of the amount of internal EGR gas contained in the mixture (total gas) in the combustion chamber”, and the “external EGR rate” is “the mixture in the combustion chamber ( The ratio of the amount of external EGR gas contained in the total gas). Details of the “heat ratio (SI rate)” will be described later.

<SPCCI combustion>
The engine 1 performs combustion in a form that combines SI (Spark Ignition) combustion and CI (Compression Ignition) 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. Combustion mode combining SI combustion and CI combustion is such that the spark plug forcibly ignites the air-fuel mixture in the combustion chamber, so that the air-fuel mixture burns by flame propagation. This is a form in which the unburned mixture is burned by self-ignition by increasing the temperature in the combustion chamber due to the pressure increase.

  By adjusting the calorific value of SI combustion, it is possible to absorb variations in temperature in the combustion chamber before the start of compression. If the start timing of SI combustion is adjusted by adjusting the ignition timing, for example, according to the temperature in the combustion chamber before the start of compression, the unburned mixture can be self-ignited at the target time.

  Since the SI combustion controls the CI combustion, a combustion mode combining SI combustion and CI combustion (SICI combustion) is hereinafter referred to as SPCCI (SPark Controlled Compression Ignition) combustion.

<Configuration of engine 1>
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 gasoline containing bioethanol or the like. The fuel of the engine 1 may be any fuel as long as it is a liquid fuel containing at least gasoline.

  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. The engine 1 further includes a piston 3, an injector 6, a spark plug 25, an intake valve 21, an exhaust valve 22, and the like.

(Piston 3)
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” is not limited to the meaning of the space when the piston 3 reaches compression top dead center. The term “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.

  As shown in FIG. 2, the lower surface of the cylinder head 13, that is, the ceiling surface of the combustion chamber 17 is composed of an inclined surface 1311 and an inclined surface 1312. The inclined surface 1311 has an upward gradient from the intake side toward the axis X2 (axis passing through the injection center of the injector 6). The inclined surface 1312 has an upward slope from the exhaust side toward the axis X2. The ceiling surface of the combustion chamber 17 has a so-called pent roof shape. A squish area 171 is formed in a region outside the cavity 31.

  The upper surface of the piston 3 is raised toward the ceiling surface of the combustion chamber 17 (may be a flat surface). A cavity 31 is formed on the upper surface of the piston 3. The cavity 31 is recessed from the upper surface of the piston 3. The cavity 31 has a shallow dish shape. The cavity 31 faces the injector 6 when the piston 3 is positioned near the compression top dead center.

  The center of the cavity 31 is shifted to the exhaust side with respect to the axis X1 (the axis passing through the center of the cylinder 11). The center of the cavity 31 coincides with the axis X2. The cavity 31 has a convex portion 311. The convex portion 311 is provided on the axis X2. The convex part 311 is substantially conical. The convex portion 311 extends upward from the bottom of the cavity 31 toward the ceiling surface of the cylinder 11.

  The cavity 31 also has a concave portion 312 provided around the convex portion 311. The recessed portion 312 is provided so as to surround the entire circumference of the protruding portion 311. The cavity 31 has a symmetrical shape with respect to the axis X2.

  The peripheral side surface of the recessed portion 312 is inclined with respect to the axis X <b> 2 from the bottom surface of the cavity 31 toward the opening of the cavity 31. The inner diameter of the cavity 31 in the recessed portion 312 gradually increases from the bottom of the cavity 31 toward the opening of the cavity 31.

  The shape of the combustion chamber 17 is not limited to the shape illustrated in FIG. For example, the shape of the cavity 31, the shape of the upper surface of the piston 3, the shape of the ceiling surface of the combustion chamber 17, and the like can be changed as appropriate.

  For example, the depth of the recessed portion 312 outside the cylinder may be reduced. In this case, EGR gas around the spark plug 25 is reduced, and flame propagation of SI combustion in SPCCI combustion is improved.

  The cavity 31 may also be asymmetric with respect to the axis X2, as indicated by the phantom line L1 in FIG. That is, the exhaust-side recess 312 is larger and deeper than the intake-side recess 312. In this case, since the fuel concentration around the spark plug 25 at the time of ignition can be increased, the SI combustion ignition in the SPCCI combustion is improved.

  The cavity 31 may also omit the convex portion 311 as indicated by a virtual line L2 in FIG. That is, the cavity 31 has a spherical shape that gradually becomes shallower from the center toward the outside in the radial direction. In this case, since it becomes difficult for the piston 3 to come into contact with the intake valve 21 and the exhaust valve 22, the degree of freedom of opening and closing control of the intake valve 21 and the exhaust valve 22 is improved. When a swirl flow is formed in the combustion chamber 17, the flow can be stabilized, so that the stratification of the air-fuel mixture is facilitated.

  The geometric compression ratio of the engine 1 is set to 13 or more and 30 or less. As described above, the SPCCI combustion uses the heat generated by the SI combustion and the pressure increase to control the CI combustion. 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 due to the self-ignition of the air-fuel mixture.

  That is, although the engine 1 performs CI combustion, the geometric compression ratio can be set relatively low. Lowering the geometric compression ratio is advantageous in reducing cooling loss and mechanical loss. As an example, the geometric compression ratio of the engine 1 is 14 to 17 (14 to 17 or less, the same applies hereinafter) in the regular specification (the fuel octane number is about 91), and the high-octane specification (the fuel octane number is about 96). In, it is good also as 15-18.

(Intake valve 21 and exhaust valve 22)
An intake port 18 is formed in the cylinder head 13 for each cylinder 11. As shown in FIG. 3, the intake port 18 has two intake ports, a first intake port 181 and a second intake port 182. The first intake port 181 and the second intake port 182 are aligned in the axial direction of the crankshaft 15, that is, the front-rear direction of the engine 1. 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 motor S-VT 23 is configured to continuously change the rotation phase of the intake camshaft within a predetermined angle range. Thereby, the valve opening timing and the valve closing timing of the intake valve 21 are continuously changed. 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 two exhaust ports, a first exhaust port 191 and a second exhaust port 192, as shown in FIG. The first exhaust port 191 and the second exhaust port 192 are arranged in the front-rear direction of the engine 1. 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 is configured to continuously change the rotation phase of the exhaust camshaft within a predetermined angle range. Thereby, the valve opening timing and the valve closing timing of the exhaust valve 22 continuously change. The exhaust valve mechanism may have a hydraulic S-VT instead of the electric S-VT.

  Although details will be described later, the engine 1 adjusts the length of the overlap period related to the opening timing of the intake valve 21 and the closing timing of the exhaust valve 22 by the intake electric S-VT 23 and the exhaust electric S-VT 24. To do. As a result, the residual gas in the combustion chamber 17 is scavenged. Further, by adjusting the length of the overlap period, an internal EGR (Exhaust Gas Recirculation) gas is introduced into the combustion chamber 17 or confined in the combustion chamber 17. In this configuration example, the intake electric S-VT 23 and the exhaust electric S-VT 24 constitute an internal EGR system as one of the state quantity setting devices. Note that the internal EGR system is not necessarily configured by S-VT.

(Injector 6)
An injector 6 is attached to the cylinder head 13 for each cylinder 11. The injector 6 is configured to inject fuel directly into the combustion chamber 17. The injector 6 is disposed in a valley portion of the pent roof where the intake-side inclined surface 1311 and the exhaust-side inclined surface 1312 intersect. As shown in FIG. 2, the injector 6 is arranged such that its injection axis is along the axis X2. The injection axis of the injector 6 coincides with the position of the convex portion 311 of the cavity 31. The injector 6 faces the cavity 31. The injection axis of the injector 6 may coincide with the central axis X1 of the cylinder 11. Even in this case, it is desirable that the injection axis of the injector 6 and the position of the convex portion 311 of the cavity 31 coincide with each other.

  Although not shown in detail, the injector 6 is constituted by a multi-injection type fuel injection valve having a plurality of injection holes. As shown by a two-dot chain line in FIG. 2, the injector 6 injects fuel so that the fuel spray spreads radially downward from the upper central portion of the combustion chamber 17 where the injection axis is located. In the present configuration example, the injector 6 has ten nozzle holes, and each nozzle hole is arranged at an equal angle in the circumferential direction.

  The axis of each nozzle hole is displaced in the circumferential direction with respect to a spark plug 25 described later, as shown in the upper diagram of FIG. That is, the spark plug 25 is sandwiched between the shafts of two adjacent nozzle holes. Thereby, it is avoided that the spray of the fuel injected from the injector 6 directly hits the spark plug 25 and wets the electrode.

  As will be described later, the injector 6 may inject fuel at the timing when the piston 3 is positioned near the compression top dead center. In that case, when the injector 6 injects the fuel, the fuel spray flows downward along the convex portion 311 of the cavity 31 while mixing with fresh air, and along the bottom surface and the peripheral side surface of the concave portion 312, the combustion chamber. From the center of 17, it spreads radially outward in the radial direction. Thereafter, the air-fuel mixture reaches the opening of the cavity 31 and flows from the radially outer side toward the center of the combustion chamber 17 along the inclined surface 1311 on the intake side and the inclined surface 1312 on the exhaust side. The injector 6 is not limited to a multi-hole type injector. The injector 6 may employ an external valve opening type injector.

  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 is configured to store 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 hole of the injector 6. The fuel supply system 61 is configured to be able to supply high pressure fuel of 30 MPa or more to the injector 6. The maximum fuel pressure of the fuel supply system 61 may be about 120 MPa, for example. 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.

(Ignition plug 25)
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 axis X1 passing through the center 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 electrode of the spark plug 25 is adjacent to the injection hole of the injector 6. The arrangement position of the spark plug 25 is not limited to the configuration example of FIG. The spark plug 25 may be disposed on the exhaust side of the axis X1. In addition, the spark plug 25 may be disposed on the axis X1, and the injector 6 may be disposed on the intake side or the exhaust side from the axis X1.

(Intake passage 40)
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 intake passage 40 is a passage through which gas introduced into the combustion chamber 17 flows. An air cleaner 41 that filters fresh air is disposed at the upstream end of the intake passage 40. A surge tank 42 is disposed near the downstream end of the intake passage 40. Although the detailed illustration is omitted, the intake passage 40 downstream of the surge tank 42 constitutes an independent passage branched 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 is configured to change the amount of fresh air introduced into the combustion chamber 17 by changing the opening of the valve. The throttle valve 43 constitutes one of state quantity setting devices.

  A supercharger 44 is also arranged in the intake passage 40 downstream of the throttle valve 43. The supercharger 44 is configured to supercharge 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, for example, a Rishorum type. The configuration of the mechanical supercharger 44 may be any configuration. The mechanical supercharger 44 may be a roots type, a vane type, or a centrifugal type. The supercharger may be an electric supercharger or a turbocharger driven by exhaust.

  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 on and off. That is, in the engine 1, the supercharger 44 can switch between supercharging the gas introduced into the combustion chamber 17 and the supercharger 44 not supercharging the gas introduced into the combustion chamber 17. It is configured to be able to.

  An intercooler 46 is disposed downstream of the supercharger 44 in the intake passage 40. The intercooler 46 is configured to cool the gas compressed in the supercharger 44. The intercooler 46 may be configured to be, for example, a water 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 (specifically, the surge tank 42) 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 changes the flow rate of the gas flowing through the bypass passage 47.

  When the supercharger 44 is turned off (that is, when the electromagnetic clutch 45 is disconnected), the air bypass valve 48 is fully opened. As a result, 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 (that is, when the electromagnetic clutch 45 is connected), the engine is operated in a supercharged state (a state in which the downstream side of the supercharger 44 is dynamically increased from atmospheric pressure). A part of the gas that has passed through the supercharger 44 flows back upstream of the supercharger 44 through the bypass passage 47. Since the reverse flow rate can be changed by changing the opening degree of the air bypass valve 48, the supercharging pressure of the gas introduced into the combustion chamber 17 can be changed. In this configuration example, the supercharger 44, the bypass passage 47, and the air bypass valve 48 constitute a supercharging system 49. The air bypass valve 48 constitutes one of state quantity setting devices.

(Swirl control valve 56)
The intake passage 40 also controls the flow of intake air introduced into the combustion chamber 17 to form a swirl flow in the combustion chamber 17 and change the strength of the swirl control valve 56 (intake flow control device). Is arranged.

  As shown in FIG. 3, 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 adjustment valve that can narrow the cross section of the secondary passage.

  When the opening of the swirl control valve 56 is small, among the first intake port 181 and the second intake port 182 aligned in the front-rear direction of the engine 1, the intake flow rate flowing into the combustion chamber 17 from the first intake port 181 is relative. And the intake flow rate flowing into the combustion chamber 17 from the second intake port 182 relatively decreases, so that the swirl flow in the combustion chamber 17 becomes stronger.

  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).

  Instead of attaching the swirl control valve 56 to the intake passage 40 or in addition to attaching the swirl control valve 56, the opening periods of the two intake valves 21 are shifted, and only one of the intake valves 21 starts the combustion chamber 17. You may employ | adopt the structure which can introduce inhalation into. Since only one of the two intake valves 21 is opened, intake air is unevenly introduced into the combustion chamber 17, so that a swirl flow can be generated in the combustion chamber 17. . Further, the swirl flow may be generated in the combustion chamber 17 by devising the shape of the intake port 18.

(Swirl style)
Here, the strength of the swirl flow in the combustion chamber 17 is defined. In this configuration example, the strength of the swirl flow in the combustion chamber 17 is represented by “swirl ratio”. The “swirl ratio” can be defined as a value obtained by dividing a value obtained by measuring and integrating the intake-flow lateral angular velocity for each valve lift by the engine angular velocity. The intake flow lateral angular velocity can be obtained based on the measurement using the rig testing apparatus shown in FIG.

  That is, in the rig testing apparatus shown in the figure, the cylinder head 13 is installed upside down on the base, and the intake port 18 is connected to an intake air supply device (not shown). A cylinder 36 is installed on the cylinder head 13, and an impulse meter 38 having a honeycomb rotor 37 is connected to the upper end of the cylinder 36. The lower surface of the impulse meter 38 is positioned at a position of 1.75 D (D is a cylinder bore diameter) from the mating surface of the cylinder head 13 and the cylinder block. Torque acting on the honeycomb-like rotor 37 is measured by an impulse meter 38 by a swirl (see an arrow in FIG. 5) generated in the cylinder 36 in response to the intake air supply, and based on this, the intake air flow lateral angular velocity can be obtained. .

  FIG. 6 shows the relationship between the opening of the swirl control valve 56 and the swirl ratio in the engine 1. FIG. 6 shows the opening degree of the swirl control valve 56 by the opening ratio with respect to the fully open section of the secondary passage 402. When the swirl control valve 56 is fully closed, the opening ratio of the secondary passage 402 becomes 0%, and when the opening of the swirl control valve 56 increases, the opening ratio of the secondary passage 402 becomes larger than 0%. When the swirl control valve 56 is fully open, the opening ratio of the secondary passage 402 is 100%.

  As illustrated in FIG. 6, in the engine 1, the swirl ratio becomes about 6 when the swirl control valve 56 is fully closed. If the swirl ratio is 4 or more, the opening degree of the swirl control valve 56 may be adjusted in a range where the opening ratio is 0 to 15%.

  If the swirl ratio is less than 4, the opening degree of the swirl control valve 56 may be adjusted within a range where the opening ratio is less than 15%. In particular, in order to ensure the fluidity of the air-fuel mixture and to control the stratification of the air-fuel mixture in the combustion chamber 17 in combination with SPCCI combustion in the operating range of the engine 1 at low and medium loads. The swirl ratio is preferably adjusted within a range of 1.5 to 3 (25% to 40% if the swirl control valve 56 is open).

(Exhaust passage 50)
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 one or more catalytic converters is disposed in the exhaust passage 50. The exhaust gas purification system has two catalytic converters in this configuration example. 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 burned gas to the intake passage 40. The upstream end of the EGR passage 52 is connected between the upstream catalytic converter and the downstream catalytic converter in the exhaust passage 50. The downstream end of the EGR passage 52 is connected to the upstream side of the supercharger 44 in the intake passage 40.

  More specifically, the downstream end of the EGR passage 52 is connected in the middle of the bypass passage 47. 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 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 is configured to cool the burned gas. An EGR valve 54 is also disposed in the EGR passage 52. The EGR valve 54 is configured to change the flow rate of burned gas flowing through the EGR passage 52. By changing the opening degree of the EGR valve 54, the recirculation amount of the cooled burned gas, that is, the external EGR gas can be changed.

  In this configuration example, the EGR system 55 includes a system (external EGR system) configured to include the EGR passage 52 and the EGR valve 54, and the above-described intake electric S-VT 23 and exhaust electric S-VT 24. System (internal EGR system). The EGR valve 54 also constitutes one of the state quantity setting devices. In the external EGR system, the EGR passage 52 is connected downstream of the catalytic converter and has an EGR cooler 53, so that burned gas having a temperature lower than that of the internal EGR system is supplied to the combustion chamber 17. Can do.

(ECU 10)
The control device of the engine 1 includes an ECU (Engine Control Unit) 10 for operating the engine 1. The ECU 10 is a controller based on a well-known microcomputer. The ECU 10 includes a central processing unit (CPU) 101 that executes a program, a memory 102 that includes, for example, a RAM (Random Access Memory) and a ROM (Read Only Memory), and stores programs and data. And an input / output bus 103 for inputting and outputting signals.

  The memory 102 stores SI rate changing means (heat quantity ratio changing means) 102a for changing the SI rate (heat quantity ratio, details will be described later) according to the operating state of the engine 1 in order to control SPCCI combustion. . The SI rate changing means 102a is composed of data such as a control program and a map used therefor, for example.

  As shown in FIGS. 1 and 3, various sensors SW <b> 1 to SW <b> 16 are connected to the ECU 10. Sensors SW1-SW16 output a detection signal to ECU10. The sensors include the following sensors.

  That is, the air flow sensor SW1 that is disposed downstream of the air cleaner 41 in the intake passage 40 and detects the flow rate of fresh air flowing through the intake passage 40, the first intake temperature sensor SW2 that detects the temperature of fresh air, and the intake passage 40, the first pressure sensor SW3 that is disposed downstream of the connection position of the EGR passage 52 and upstream of the supercharger 44 and detects the pressure of the gas flowing into the supercharger 44, and supercharging in the intake passage 40 The second intake air temperature sensor SW4, which is disposed downstream of the machine 44 and upstream of the connection position of the bypass passage 47 and detects the temperature of the gas flowing out from the supercharger 44, is attached to the surge tank 42, and A second pressure sensor SW5 for detecting the pressure of the gas downstream of the feeder 44, attached to the cylinder head 13 corresponding to each cylinder 11, and each A finger pressure sensor SW6 that detects the pressure in the firing chamber 17 and the exhaust passage 50 and is disposed upstream of the catalytic converter 51 in the exhaust passage 50 and an exhaust temperature sensor SW7 that detects the temperature of the exhaust gas discharged from the combustion chamber 17 And a linear O2 sensor SW8 that detects the oxygen concentration in the exhaust gas, a lambda O2 sensor SW9 that is disposed downstream of the catalytic converter 51 in the exhaust passage 50 and detects the oxygen concentration in the exhaust gas, and is attached to the engine 1. Further, a water temperature sensor SW10 that detects the temperature of the cooling water, a crank angle sensor SW11 that is attached to the engine 1 and that detects the rotation angle of the crankshaft 15, is attached to the accelerator pedal mechanism, and corresponds to the operation amount of the accelerator pedal. Accelerator position sensor SW12 for detecting the accelerator position, engine 1, an intake cam angle sensor SW13 that detects the rotation angle of the intake camshaft, an exhaust cam angle sensor SW14 that is attached to the engine 1 and detects the rotation angle of the exhaust camshaft, and is disposed in the EGR passage 52. An EGR differential pressure sensor SW15 that detects a differential pressure upstream and downstream of the EGR valve 54, and a fuel pressure sensor SW16 that is attached to the common rail 64 of the fuel supply system 61 and detects the pressure of the fuel supplied to the injector 6. .

  The ECU 10 determines the operating state of the engine 1 based on these detection signals and calculates the control amount of each device. The ECU 100 sends a control signal related to the calculated control amount to the electromagnetic clutch 45 of the injector 6, spark plug 25, intake electric VVT 23, exhaust electric VVT 24, fuel supply system 61, throttle valve 43, EGR valve 54, and supercharger 44. , And output to the air bypass valve 48 and the swirl control valve 56.

  For example, the ECU 10 changes the opening of the air bypass valve 48 based on the differential pressure across the supercharger 44 obtained from the detection signals of the first pressure sensor SW3 and the second pressure sensor SW5, thereby increasing the supercharging pressure. change. Further, the ECU 10 changes the opening degree of the EGR valve 54 based on the differential pressure across the EGR valve 54 obtained from the detection signal of the EGR differential pressure sensor SW15, so that the external EGR gas introduced into the combustion chamber 17 is changed. Change the amount. Details of control of the engine 1 by the ECU 10 will be described later.

<Engine operating range>
FIG. 7A illustrates a first configuration example (first operation region map 700) of the operation region map of the engine 1. The operation region map 700 is determined by the load and the rotational speed, and is roughly divided into the following four regions with respect to the load level and the rotational speed level.

(A) Low load region including idle operation (B) Medium load region between low load region (A) and next high load region (C) (C) High load region including fully open load (D) Low load High rotational speed region having a higher rotational speed than the region (A), medium load region (B), and high load region (C)

  The engine 1 performs SPCCI combustion in the medium load region (B) mainly for the purpose of improving fuel consumption and exhaust gas performance. Hereinafter, the combustion modes in each of the low load region (A), the medium load region (B), and the high load region (C) will be described in detail.

(Low load area)
When the operating state of the engine 1 is in the low load region (A), the fuel injection amount is small. Therefore, the amount of heat generated when the air-fuel mixture burns in the combustion chamber 17 is small, and the temperature of the combustion chamber 17 is lowered. Further, since the temperature of the exhaust gas is lowered, even if the internal EGR gas is introduced into the combustion chamber 17 as will be described later, the temperature of the combustion chamber 17 is increased to such an extent that self-ignition is possible stably. Is difficult.

  The combustion mode when the operating state of the engine 1 is in the low load region (A) is SI combustion in which the air-fuel mixture is combusted by flame propagation when the spark plug 25 ignites the air-fuel mixture in the combustion chamber 17. . Hereinafter, the combustion mode in the low load region (A) may be referred to as low load SI combustion.

  When the operating state of the engine 1 is in the low load region (A), the air-fuel ratio (A / F) of the air-fuel mixture is the stoichiometric air-fuel ratio (A / F≈14.7). In the following description, the air-fuel ratio of the air-fuel mixture, the excess air ratio λ, and G / F mean values at the ignition timing.

  When the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio, the exhaust gas performance of the engine 1 is improved because the three-way catalyst can purify the exhaust gas discharged from the combustion chamber 17. 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.

  In order to improve the fuel consumption performance of the engine 1, the EGR system 55 introduces EGR gas into the combustion chamber 17 when the operating state of the engine 1 is in the low load region (A). The G / F of the air-fuel mixture, that is, the mass ratio of the total gas and fuel in the combustion chamber 17 is set to 18 or more and 30 or less. The G / F of the air-fuel mixture may be set to 18 or more and 50 or less. The mixture is EGR lean. The dilution ratio of the mixture is high.

  If the G / F of the air-fuel mixture is set to 25, for example, SI combustion can be performed stably in the low load region (A) without the air-fuel mixture reaching self-ignition. In the low load region (A), the G / F of the air-fuel mixture is kept constant regardless of the load level of the engine 1. By doing so, SI combustion is stabilized throughout the low load region. Further, the fuel efficiency of the engine 1 is improved and the exhaust gas performance is improved.

  When the operating state of the engine 1 is in the low load region (A), since the amount of fuel is small, in order to set the λ of the air-fuel mixture to 1.0 ± 0.2 and G / F to 18 or more and 50 or less, The filling amount of the gas introduced into the combustion chamber 17 must be less than 100%. Specifically, the engine 1 executes throttling for changing the opening degree of the throttle valve 43 and / or a mirror cycle for delaying the closing timing of the intake valve 21 after the intake bottom dead center.

  In the engine 1 employing the operation region map 700, when the operation state is in the low load region (A), the opening degree of the swirl control valve 56 is substantially fully opened. Accordingly, almost no swirl flow is generated in the combustion chamber 17.

  In the low load region (A), in the low load and low speed region, the combustion temperature of the air-fuel mixture and the exhaust gas temperature may be increased by further reducing the gas filling amount. This is advantageous in maintaining the catalytic converter 51 in an active state.

(Medium load area)
When the operating state of the engine 1 is in the medium load region (B), the fuel injection amount increases. Since the temperature of the combustion chamber 17 becomes high, it becomes possible to perform self-ignition stably. In order to improve fuel consumption and exhaust gas performance, the engine 1 performs CI combustion in the medium load region (B).

  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 in the medium load region (B).

In the middle load region (B), the engine 1 is in a state in the combustion chamber 17 such that the λ of the mixture is 1.0 ± 0.2 and the G / F of the mixture is 18 or more and 50 or less. To do. The required temperature T _IG in the combustion chamber 17 at the ignition timing is 570 to 800 K, the required pressure P _IG in the combustion chamber 17 at the ignition timing is 400 to 920 kPa, and the turbulent energy in the fuel chamber 17 is 17 to 40 m ² / s ² .

  In the engine 1 employing the operation region map 700, when the operation state is in the medium load region (B), the opening degree of the swirl control valve 56 is substantially fully opened. Accordingly, almost no swirl flow is generated in the combustion chamber 17.

  By controlling the timing of self-ignition with high accuracy, an increase in combustion noise can be avoided when the operating state of the engine 1 is in the medium load region (B). Further, by performing the CI combustion with the dilution ratio of the air-fuel mixture as high as possible, the fuel efficiency performance of the engine 1 can be enhanced. Furthermore, by setting λ of the air-fuel mixture to 1.0 ± 0.2, it becomes possible to purify the exhaust gas by the three-way catalyst, so that the exhaust gas performance of the engine 1 becomes good.

  As described above, in the low load region (A), the G / F of the air-fuel mixture is set to 18 to 50 (for example, 25), and λ of the air-fuel mixture is set to 1.0 ± 0.2. The state quantity in the combustion chamber 17 does not fluctuate greatly between when the operating state of the engine 1 is in the low load region (A) and when it is in the medium load region (B). Therefore, the robustness of the control of the engine 1 against the change in the load of the engine 1 is enhanced.

  When the operating state of the engine 1 is in the medium load region (B), the amount of fuel increases, unlike in the low load region (A), so the filling amount of the gas introduced into the combustion chamber 17 is changed. There is no need to do. The opening degree of the throttle valve 43 is fully open.

  When the load of the engine 1 is increased and the fuel amount is further increased, the natural intake air can be reduced by setting the λ of the mixture to 1.0 ± 0.2 and the G / F of the mixture to 18 to 50. If so, the amount of gas introduced into the combustion chamber 17 is insufficient. Therefore, the supercharger 44 supercharges the gas introduced into the combustion chamber 17 in a region where the load is higher than the predetermined load in the medium load region (B).

  The medium load region (B) is a region that is higher than the predetermined load, and is a first medium load region (B1) that performs supercharging, and a region that is below the predetermined load and that does not perform supercharging. It is divided into a medium load region (B2). The predetermined load is, for example, a ½ load. The second medium load region (B2) is a region having a lower load than the first medium load region (B1). Hereinafter, the combustion mode in the first medium load region (B1) may be referred to as supercharging SPCCI combustion, and the combustion mode in the second medium load region (B2) may be referred to as non-supercharging SPCCI combustion.

  In the second medium load region (B2) where supercharging is not performed, as the amount of fuel increases, new air introduced into the combustion chamber 17 increases while EGR gas decreases. The G / F of the air-fuel mixture decreases as the load on the engine 1 increases. Since the opening degree of the throttle valve 43 is fully opened, the engine 1 changes the amount of fresh air introduced into the combustion chamber 17 by changing the amount of EGR gas introduced into the combustion chamber 17. . In the second intermediate load region (B2), for example, the state quantity in the combustion chamber 17 is substantially constant when λ of the air-fuel mixture is 1.0, while the G / F of the air-fuel mixture is changed within the range of 25 to 28. The

  In contrast, in the first medium load region (B1) where supercharging is performed, the engine 1 increases both fresh air and EGR gas introduced into the combustion chamber 17 as the amount of fuel increases. The G / F of the air-fuel mixture is constant even when the load on the engine 1 increases. In the first medium load region (B1), the state quantity in the combustion chamber 17 is, for example, λ of the air-fuel mixture becomes substantially constant at 1.0, and the G / F of the air-fuel mixture is constant at 25.

(High load area)
The combustion mode when the operating state of the engine 1 is in the high load region is SI combustion. This is because priority is given to reliably avoiding combustion noise. Hereinafter, the combustion mode in the high load region may be referred to as high load SI combustion.

  When the operating state of the engine 1 is in the high load region (C), λ of the air-fuel mixture is 1.0 ± 0.2. The G / F of the air-fuel mixture is basically set to 18 or more and 30 or less. The G / F of the air-fuel mixture may be set to 18 or more and 50 or less. In the high load region (C), the opening degree of the throttle valve 43 is fully open, and the supercharger 44 performs supercharging.

  In the high load region (C), the engine 1 reduces the amount of EGR gas as the load increases. The G / F of the air-fuel mixture decreases as the load on the engine 1 increases. Since the amount of fresh air introduced into the combustion chamber 17 is increased by the amount of EGR gas reduced, the amount of fuel can be increased. This is advantageous in increasing the maximum output of the engine 1.

  In the engine 1 employing the operation region map 700, when the operation state is in the high load region (C), the opening degree of the swirl control valve 56 is substantially fully opened. Accordingly, almost no swirl flow is generated in the combustion chamber 17.

  The state quantity in the combustion chamber 17 does not fluctuate greatly between when the operating state of the engine 1 is in the high load region (C) and when it is in the medium load region (B). The robustness of the control of the engine 1 against the change of the load of the engine 1 is increased.

  As described above, the engine 1 performs SI combustion in the high load region (C), but there is a problem that abnormal combustion such as pre-ignition and knocking is likely to occur.

  Therefore, the engine 1 is configured to avoid abnormal combustion by devising the form of fuel injection in the high load region (C). Specifically, the ECU 10 injects fuel into the combustion chamber 17 at a high fuel pressure of 30 MPa or more and at a timing within a period from the latter stage of the compression stroke to the early stage of the expansion stroke (hereinafter, this period is referred to as a retard period). Thus, a control signal is output to the fuel supply system 61 and the injector 6. The ECU 10 also outputs a control signal to the spark plug 25 so that the air-fuel mixture is ignited at a timing near the compression top dead center after fuel injection. In the following description, injecting fuel into the combustion chamber 17 at a high fuel pressure and at a timing within the retard period is referred to as high-pressure retarded injection.

  High pressure retarded injection avoids abnormal combustion by shortening the time during which the air-fuel mixture reacts. That is, the reaction time of the air-fuel mixture includes (1) a period during which the injector 6 injects fuel (that is, an injection period), and (2) after the fuel injection is completed, This is a time obtained by adding the period until formation (that is, the mixture formation period) and (3) the period until SI combustion started by ignition ends ((3) combustion period).

  When fuel is injected into the combustion chamber 17 at a high fuel pressure, the injection period and the mixture formation period become shorter. When the injection period and the air-fuel mixture formation period are shortened, the timing for starting fuel injection can be made closer to the ignition timing. In the high pressure retarded injection, since the fuel is injected into the combustion chamber 17 at a high pressure, the fuel is injected at the timing within the retard period from the latter stage of the compression stroke to the early stage of the expansion stroke.

  When fuel is injected into the combustion chamber 17 at a high fuel pressure, the turbulent energy in the combustion chamber 17 increases. When the fuel injection timing is brought close to the compression top dead center, SI combustion can be started with high turbulent energy in the combustion chamber 17. As a result, the combustion period is shortened.

  The high pressure retarded injection can shorten the injection period, the mixture formation period, and the combustion period. Compared with the case where fuel is injected into the combustion chamber 17 during the intake stroke, the high-pressure retarded injection can greatly shorten the time for the air-fuel mixture to react. In the high pressure retarded injection, the time for which the air-fuel mixture reacts is shortened, so that abnormal combustion can be avoided.

  In the technical field of engine control, in order to avoid abnormal combustion, the ignition timing is conventionally retarded. However, if the ignition timing is delayed, the fuel consumption performance decreases. The high pressure retarded injection need not retard the ignition timing. By using high-pressure retarded injection, fuel efficiency is improved.

  If the fuel pressure is, for example, 30 MPa or more, the injection period, the mixture formation period, and the combustion period can be effectively shortened. The fuel pressure is preferably set as appropriate according to the properties of the fuel. As an example, the upper limit value of the fuel pressure may be 120 MPa.

  Here, when the rotation speed of the engine 1 is low, it takes a long time for the crank angle to change by the same angle. Therefore, shortening the reaction time of the air-fuel mixture by high-pressure retarded injection is to avoid abnormal combustion. Is particularly effective. On the other hand, when the rotational speed of the engine 1 is increased, the time when the crank angle changes by the same angle is shortened. For this reason, shortening the reaction time of the air-fuel mixture is not so effective in avoiding abnormal combustion.

  The high-pressure retarded injection also injects fuel into the combustion chamber 17 only after the compression top dead center is reached. Therefore, in the compression stroke, in the combustion chamber 17, a gas that does not contain fuel, in other words, a specific heat ratio. High gas is compressed. If high-pressure retarded injection is performed when the rotational speed of the engine 1 is high, the temperature in the combustion chamber 17 at the compression top dead center, that is, the compression end temperature becomes high. An increase in the compression end temperature may cause abnormal combustion such as knocking.

  Therefore, in the engine 1, the high load region (C) is divided into the first high load region (C1) on the low rotation side and the second high load region (C2) having a higher rotational speed than the first high load region (C1). ). The first high load region (C1) may include a low rotation region and a medium rotation region when the high load region (C) is divided into three regions of low rotation, medium rotation, and high rotation. The second high load region (C2) may include a high rotation region when the high load region (C) is divided into three regions of low rotation, medium rotation, and high rotation.

  In the first high load region (C1), the injector 6 receives the control signal of the ECU 10 and performs the above-described high pressure retarded injection. In the second high load region (C2), the injector 6 receives a control signal from the ECU 10 and injects fuel at a predetermined timing during the intake stroke. The fuel injection performed during the intake stroke does not require high fuel pressure. The ECU 10 outputs a control signal to the fuel supply system 61 so that the fuel pressure is lower than the fuel pressure of the high pressure retarded injection (for example, the fuel pressure is less than 40 MPa). By reducing the fuel pressure, the mechanical resistance loss of the engine 1 is reduced, which is advantageous for improving fuel consumption.

  By injecting fuel into the combustion chamber 17 during the intake stroke, the specific heat ratio of the gas in the combustion chamber 17 is lowered, so that the compression end temperature is lowered. Since the compression end temperature becomes low, the engine 1 can avoid abnormal combustion. Since it is not necessary to retard the ignition timing in order to avoid abnormal combustion, in the second high load region (C2), the spark plug 25 is compressed top dead center as in the first high load region (C1). The mixture is ignited at a nearby timing.

  In the first high load region (C1), the engine 1 can perform stable SI combustion because the air-fuel mixture does not reach self-ignition by high-pressure retarded injection. In the second high load region (C2), the air-fuel mixture does not reach self-ignition due to fuel injection during the intake stroke, and therefore the engine 1 can perform stable SI combustion.

<SPCCI combustion>
Next, the aforementioned SPCCI combustion will be described in more detail. The upper diagram of FIG. 8 shows a waveform 801 illustrating the change in the heat generation rate with respect to the crank angle in SPCCI combustion. When the ignition plug 25 ignites the air-fuel mixture in the vicinity of the compression top dead center, more precisely at a predetermined timing before the compression top dead center, combustion by flame propagation starts. Heat generation during SI combustion is milder than heat generation during CI combustion. Therefore, the waveform of the heat generation rate has a relatively small slope. Although not shown, the pressure fluctuation (dp / dθ) in the combustion chamber 17 during SI combustion is also gentler than that during CI combustion.

  When the temperature and pressure in the combustion chamber 17 are increased by SI combustion, the unburned mixture is self-ignited. In the example of the waveform 801, the slope of the heat generation rate waveform changes from small to large near the compression top dead center. That is, the heat generation rate waveform has an inflection point 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, the piston 3 is lowered by motoring. It is avoided that the inclination of the waveform of the heat generation rate due to CI combustion becomes too large. Dp / dθ during CI combustion also becomes relatively gentle.

  dp / dθ can be used as an index representing combustion noise. However, since SPCCI combustion can reduce dp / dθ as described above, it is possible to avoid excessive combustion noise. . Combustion noise can be suppressed below an acceptable level.

  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. In other words, SPCCI combustion can bring the combustion end time during the expansion stroke closer to the compression top dead center. The SPCCI combustion is more advantageous for improving the fuel efficiency of the engine 1 than the SI combustion.

  Therefore, SPCCI combustion can achieve both prevention of combustion noise and improvement in fuel efficiency.

(SI rate: calorie ratio)
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. The SI rate may be the ratio of the amount of heat generated by SI combustion to the amount of heat generated by SPCCI combustion. For example, in the waveform 801, the SI rate can be expressed by SI rate = (SI combustion area) / (SPCCI combustion area). In the waveform 801, the SI rate may be referred to as an SI fuel rate in the meaning of the ratio of fuel that burns by SI combustion.

  The SI rate is a ratio between SI combustion and CI combustion in SPCCI combustion combining SI combustion and CI combustion. 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 is not limited to the above-described definition. Various definitions can be considered for the SI rate. For example, the SI rate may be the ratio of the amount of heat generated by SI combustion to the amount of heat generated by CI combustion. That is, in FIG. 5, SI rate = (SI combustion area) / (CI combustion area).

Further, in SPCCI combustion, the waveform of the heat generation rate has an inflection point at the timing when CI combustion starts. Therefore, as indicated by reference numeral 802 in the middle diagram of FIG. 8, the inflection point in the heat generation rate waveform is used as a boundary, the range on the advance side of the boundary is SI combustion, and the range on the retard side is CI combustion. Good. In this case, as shown by hatching the waveform 802, the SI rate is calculated from the area Q _{SI in} the range on the advance side of the boundary and the area Q _{CI in} the range on the retard side from the boundary, and the SI rate = Q _SI / (Q _SI + Q _CI ) or SI rate = Q _SI / Q _CI . In addition, the SI rate may be defined on the basis of a part of the area rather than the entire area on the advance side from the boundary and a part of the area on the retard side from the boundary.

In addition, the SI rate is not defined based on the heat generation, but the SI rate = Δθ _SI / (Δθ) from the crank angle Δθ _{SI in} the range on the advance side from the boundary and the crank angle Δθ _{CI in} the range on the retard side. _SI + [Delta] [theta] _CI) may be used as the, or as a SI index = Δθ _SI / Δθ _CI.

Furthermore, the SI rate = ΔP _SI / (ΔP _SI + ΔP _CI ) may be calculated from the peak ΔP _SI of the heat generation rate in the range on the advance side from the boundary and the peak ΔP _CI of the heat generation rate in the range on the retard side. , SI rate = ΔP _SI / ΔP _CI .

In addition, the SI rate = φ _SI / (φ _SI + φ _CI ) may be calculated from the slope φ _SI of the heat generation rate in the advance angle range from the boundary and the slope φ _CI of the heat generation rate in the retard angle range. The SI rate may be φ _SI / φ _CI .

  Also, here, based on the heat generation rate waveform, the area (that is, the amount of heat generation), the length of the horizontal axis (that is, the size of the crank angle), the length of the vertical axis (that is, the heat generation amount) The SI rate is defined from the magnitude of the generation rate) or the slope (that is, the rate of change of the heat generation rate). Although illustration is omitted, based on the waveform of the pressure (P) in the combustion chamber 17, the SI rate is similarly defined from the area, the length of the horizontal axis, the length of the vertical axis, or the slope. Also good.

Further, in SPCCI combustion, the inflection point of the combustion waveform related to the heat generation rate or pressure does not always appear clearly. The following definition may be used as the definition of the SI rate that is not based on the inflection point. That is, as indicated by reference numeral 803 in the lower diagram of FIG. 8, in the combustion waveform, the range on the advance side from the compression top dead center (TDC) is SI combustion, and the range on the retard side from the compression top dead center is CI. It is good also as combustion. In addition, as described above, the area (Q _SI , Q _CI ), the length of the horizontal axis (Δθ _SI , Δθ _CI ), the length of the vertical axis (ΔP _SI , ΔP _CI ), or the slope (φ _SI , Φ _CI ), the SI rate may be defined.

  Further, the SI rate may be defined not based on the combustion waveform actually performed in the combustion chamber 17 but based on the fuel amount. As will be described later, in the middle load region where the SPCCI combustion is performed, there is a case where split injection including the front injection and the rear injection is performed. The fuel injected into the combustion chamber 17 by the post-injection is not diffused in the combustion chamber 17 and is positioned in the vicinity of the spark plug 25 because the time from injection to ignition is short. Therefore, the fuel injected into the combustion chamber 17 by the post-stage injection mainly burns by SI combustion.

On the other hand, the fuel injected into the combustion chamber 17 by the pre-stage injection mainly burns by CI combustion. Therefore, it is possible to define the SI rate based on the fuel amount (m ₁ ) injected by the front injection and the fuel amount (m ₂ ) injected by the rear injection. That is, SI rate = m ₂ / (m ₁ + m ₂ ) may be set, or SI rate = m ₂ / m ₁ may be set.

(Stabilization of SPCCI combustion)
In order to properly perform SPCCI combustion, SI combustion must be stabilized. If the SI combustion is unstable, the entire combustion including the CI combustion is not stabilized.

  One of the factors related to the stability of SI combustion is turbulent combustion speed. When the turbulent combustion rate is high, SI combustion is stabilized. The turbulent combustion speed is affected by the air-fuel ratio (or excess air ratio λ) of the mixture, the G / F of the mixture, the temperature and pressure in the combustion chamber, and the turbulent energy in the combustion chamber. .

According to the study by the inventors of the present application, the λ of the mixture is 1.0 ± 0.2, the G / F range of the mixture is 18 or more and 30 or less, and the range of the required temperature TIg in the combustion chamber at the ignition timing. Is 570 to 800 K, the required pressure PIg in the combustion chamber at the ignition timing is 400 to 920 kPa, and the turbulent energy in the combustion chamber is 17 to 40 m ² / s ² , the SI combustion is stabilized I confirmed that I can do it.

  Furthermore, it was also confirmed that the G / F range of the air-fuel mixture can be expanded to a range of more than 30 to 50 or less by stratifying the air-fuel mixture in the combustion chamber.

  That is, the SI combustion in the SPCCI combustion is the combustion of the air-fuel mixture that is ignited by the spark plug 25. The air-fuel mixture in the vicinity of the spark plug 25 is mainly burned by SI combustion. For this reason, for example, a swirl flow is utilized to control the state of spray of fuel moving in the combustion chamber 17. By doing so, even if the G / F range in the combustion chamber exceeds 30 at the timing of ignition, the G / F of the air-fuel mixture in the vicinity of the spark plug 25 is, for example, 14 or more and 22 or less from the spark plug 25. It can be made relatively smaller than the surrounding air-fuel mixture (stratification).

  Thereby, even if it is the diluted air-fuel | gaseous mixture which is advantageous for the improvement of a fuel consumption performance, SI combustion can be stabilized and SPCCI combustion can be performed appropriately.

<Engine operation control for load direction>
The engine 1 switches between SI combustion and SPCCI combustion according to the operation state by adopting the operation region map 700. The engine 1 also changes the SI rate according to the operating state of the engine 1. Thereby, the engine 1 is compatible with suppressing the generation of combustion noise and improving the fuel consumption.

  FIG. 9 shows the change in the SI rate with respect to the load of the engine 1, the change in the state quantity in the combustion chamber 17, the change in the valve opening period of the intake valve 21 and the valve opening period of the exhaust valve 22, and the fuel The change of injection timing and ignition timing is illustrated. Hereinafter, the operation control of the engine 1 will be described on the assumption that the load of the engine 1 gradually increases at a predetermined rotational speed.

(Low load area (low load SI combustion))
In the low load region (A), the engine 1 performs low load SI combustion. When the operating state of the engine 1 is in the low load region (A), the SI rate is constant at 100%.

  In the low load region (A), as described above, the G / F of the air-fuel mixture is made constant between 18 and 50. The engine 1 introduces fresh air and burned gas in an amount corresponding to the amount of fuel into the combustion chamber 17. As described above, the amount of fresh air introduced is changed by throttling and / or mirror cycles. Since the dilution rate is high, the temperature in the combustion chamber 17 is increased in order to stabilize the SI combustion. The engine 1 introduces internal EGR gas into the combustion chamber 17 in the low load region (A).

  The internal EGR gas is introduced into the combustion chamber 17 by providing a negative overlap period in which both the intake valve 21 and the exhaust valve 22 are closed across the exhaust top dead center (that is, the burned gas is introduced into the combustion chamber). 17). The internal EGR gas amount can be changed by changing the opening timing of the intake valve 21 by the intake electric S-VT 23 and changing the closing timing of the exhaust valve 22 by the exhaust electric S-VT 24. This is done by appropriately setting the length of the lap period.

  The internal EGR gas may be introduced into the combustion chamber 17 by providing a positive overlap period in which both the intake valve 21 and the exhaust valve 22 are opened.

  That is, a period in which the valve opening period of the intake valve 21 and the valve opening period of the exhaust valve 22 overlap is provided. By doing so, part of the high-temperature burned gas in the combustion chamber 17 flows into the intake port 18 communicated with the combustion chamber 17 by opening the intake valve 21. The burned gas that has flowed into the intake port 18 is again introduced into the combustion chamber 17 during the intake stroke. The amount of internal EGR gas can be changed by changing the length of the positive overlap period and the timing of the positive overlap period.

  The introduction of the internal EGR gas by setting the positive overlap period can advance the valve opening period of the intake valve 21 relative to the introduction of the internal EGR gas by setting the negative overlap period. If the opening period of the intake valve 21 is advanced, the effective compression ratio can be increased and the temperature in the combustion chamber 17 can be increased. Therefore, it is possible to set the geometric compression ratio of the engine 1 low while ensuring the stability of SPCCI combustion as much as the temperature rise is obtained, and to reduce cooling loss, mechanical loss, pump loss, etc. There is an advantage that can be achieved.

  Also, the introduction of the internal EGR gas by setting the positive overlap period lowers the temperature of the combustion chamber than the introduction of the internal EGR gas by setting the negative overlap period in which the high-temperature burned gas is confined as it is in the combustion chamber. it can. Thereby, self-ignition in SPCCI combustion can be slowed down, and excessive CI combustion can be suppressed.

  In the low load region (A), the filling amount introduced into the combustion chamber 17 is changed to less than 100%. As the amount of fuel increases, the amount of fresh air introduced into the combustion chamber 17 and the amount of internal EGR gas gradually increase. The EGR rate in the low load region (A) is, for example, 40%.

  The injector 6 injects fuel into the combustion chamber 17 during the intake stroke. In the combustion chamber 17, a homogeneous air-fuel mixture is formed in which the excess air ratio λ is 1.0 ± 0.2 and the G / F is 18 to 50. The excess air ratio λ is preferably 1.0 to 1.2. When the spark plug 25 ignites the air-fuel mixture at a predetermined timing before the compression top dead center, the air-fuel mixture burns by flame propagation without reaching self-ignition.

(Second medium load region (non-supercharging SPCCI combustion))
When the load of the engine 1 becomes high and the operation state enters the second medium load region (B2), the engine 1 switches from the low load SI combustion to the non-supercharged SPCCI combustion. The SI rate is less than 100%. As the load on the engine 1 increases, the amount of fuel increases. When the load is low in the second medium load region (B2), the proportion of CI combustion is increased as the fuel amount increases. The SI rate gradually decreases as the load on the engine 1 increases. In the example of FIG. 9, the SI rate decreases to a predetermined value (minimum value) of 50% or less.

  Since the amount of fuel increases, the combustion temperature becomes higher in the second medium load region (B2). If the temperature in the combustion chamber 17 becomes too high, heat generation at the start of CI combustion becomes intense. If it becomes so, combustion noise will increase.

  Therefore, in the second medium load region (B2), the internal EGR gas and the external EGR gas are changed against the change in the load of the engine 1 in order to change the temperature of the combustion chamber 17 before the start of compression. And change the ratio. That is, as the load on the engine 1 increases, the hot internal EGR gas is gradually reduced and the cooled external EGR gas is gradually increased. The negative overlap period is changed from the maximum to zero in the second medium load region (B2) as the load increases. The internal EGR gas becomes zero when the load becomes highest in the second medium load region (B2).

  The same applies to the case where the internal EGR gas is introduced into the combustion chamber 17 by providing a positive overlap period of the intake valve 21 and the exhaust valve 22. As a result of adjusting the temperature in the combustion chamber 17 by adjusting the overlap period, the SI rate of SPCCI combustion can be adjusted.

  The opening degree of the EGR valve 54 is changed so that the external EGR gas increases as the load increases in the second middle load region (B2). The amount of the external EGR gas introduced into the combustion chamber 17 is changed, for example, between 0 to 30% when expressed in terms of the EGR rate. In the second medium load region (B2), the EGR gas is replaced from the internal EGR gas to the external EGR gas as the load on the engine 1 increases. Since the temperature in the combustion chamber 17 also changes by adjusting the EGR rate, the SI rate of SPCCI combustion can be adjusted.

  The amount of EGR gas introduced into the combustion chamber 17 is continuous between the low load region (A) and the second medium load region (B2). In the low load region in the second medium load region (B2), a large amount of internal EGR gas is introduced into the combustion chamber 17 as in the low load region (A). Since the temperature in the combustion chamber 17 becomes high, the air-fuel mixture surely self-ignites when the load on the engine 1 is low. The external EGR gas is introduced into the combustion chamber 17 in the high load region in the second medium load region (B2). Since the temperature in the combustion chamber 17 becomes low, the combustion noise accompanying CI combustion can be suppressed when the load of the engine 1 is high.

  In the second medium load region (B2), the filling amount introduced into the combustion chamber 17 is set to 100%. The opening degree of the throttle valve 43 is fully open. By changing the amount of EGR gas that is a combination of the internal EGR gas and the external EGR gas, the amount of fresh air introduced into the combustion chamber 17 is changed to an amount corresponding to the fuel amount.

  In the non-supercharged SPCCI combustion, the timing of self-ignition is advanced as the CI combustion ratio increases. If the timing of self-ignition becomes earlier than the compression top dead center, heat generation when CI combustion starts becomes intense. If it becomes so, combustion noise will increase. Therefore, when the load on the engine 1 reaches the predetermined load L1, the engine 1 gradually increases the SI rate as the load on the engine 1 increases.

  That is, the engine 1 increases the rate of SI combustion as the fuel amount increases. Specifically, as shown in the upper diagram of FIG. 10, in the non-supercharging SPCCI combustion, the ignition timing is gradually advanced as the fuel amount increases. As described above, since the temperature in the combustion chamber 17 is changed by reducing the amount of internal EGR gas introduced and increasing the amount of external EGR gas introduced, the SI rate increases as the fuel amount increases. Even if the temperature is increased, the temperature rise at the compression top dead center can be suppressed. The slope of the heat generation rate of SI combustion hardly changes even when the load increases. If the ignition timing is advanced, the amount of heat generated by SI combustion increases as the SI combustion starts earlier.

  As a result of suppressing the temperature rise in the combustion chamber 17 due to SI combustion, the unburned air-fuel mixture self-ignites at a timing after the compression top dead center. The heat generation by the CI combustion is almost the same even if the load of the engine 1 is high because the heat generation amount of the SI combustion is increased. Therefore, it is possible to avoid an increase in combustion noise by setting the SI rate gradually higher in accordance with the load on the engine 1 becoming higher. Note that the combustion center of gravity of the non-supercharged SPCCI combustion is retarded as the load increases.

  In the second intermediate load region (B2), the injector 6 injects fuel into the combustion chamber 17 in two steps of the front injection and the rear injection during the compression stroke. The front-stage injection injects fuel at a timing away from the ignition timing, and the rear-stage injection injects fuel at a timing close to the ignition timing.

(First medium load region (supercharged SPCCI combustion))
When the load on the engine 1 further increases and the operating state of the engine 1 enters the first medium load region (B1), the supercharger 44 supercharges fresh air and external EGR gas. Both the amount of fresh air introduced into the combustion chamber 17 and the amount of external EGR gas increase as the load on the engine 1 increases. The amount of external EGR gas introduced into the combustion chamber 17 is, for example, 30% in terms of the EGR rate. The EGR rate is constant regardless of the load level of the engine 1. Accordingly, the G / F of the air-fuel mixture is also constant regardless of the load of the engine 1. The amount of EGR gas introduced into the combustion chamber 17 is continuous between the second medium load region (B2) and the first medium load region (B1).

  The SI rate is a predetermined value less than 100%, and is constant with respect to the load of the engine 1. The SI rate of the second medium load region (B2), in particular, the SI rate that is higher than the predetermined load L1 and gradually increases as the load of the engine 1 increases, and the SI rate of the first medium load region (B1). When compared, the SI rate in the first medium load region (B1) where the load of the engine 1 is high is higher than the SI rate in the second medium load region (B2). The SI rate is continuous at the boundary between the first medium load region (B1) and the second medium load region (B2).

  Here, in the first medium load region (B1), the SI rate may be slightly changed in response to the load of the engine 1 changing. The change rate of the SI rate with respect to the load change of the engine 1 in the first medium load region (B1) may be smaller than the change rate of the SI rate on the high load side of the second medium load region (B2).

  As shown in the lower diagram of FIG. 10, in the supercharged SPCCI combustion, the ignition timing is gradually advanced as the fuel amount increases. As described above, since the amount of fresh air and EGR gas introduced into the combustion chamber 17 is increased by supercharging, the heat capacity is large. Even if the amount of fuel increases, the temperature rise in the combustion chamber 17 due to SI combustion can be suppressed. The waveform of the heat release rate of supercharged SPCCI combustion increases in a similar manner as the load increases.

  In other words, the amount of heat generated by SI combustion increases while the slope of the heat generation rate of SI combustion hardly changes. The unburned mixture self-ignites at approximately the same timing after compression top dead center. The amount of heat generated by CI combustion increases as the load on the engine 1 increases. As a result, in the first intermediate load region (B1), both the heat generation amount of SI combustion and the heat generation amount of CI combustion increase, so the SI rate becomes constant with respect to the load of the engine 1. When the peak of heat generation of CI combustion increases, the combustion noise increases. However, since the load of the engine 1 is relatively high in the first medium load region (B1), a certain level of combustion noise may be allowed. it can. Note that the center of combustion of supercharged SPCCI combustion is retarded as the load increases.

  In the first intermediate load region (B1), an overlap period (positive overlap period) in which both the intake valve 21 and the exhaust valve 22 are opened is provided with the exhaust top dead center in between. The burned gas remaining in the combustion chamber 17 is scavenged by the supercharging pressure. Thereby, since the temperature in the combustion chamber 17 becomes low, it can suppress that abnormal combustion generate | occur | produces when the load of the engine 1 is comparatively high. Further, by lowering the temperature in the combustion chamber 17, the self-ignition timing can be set to an appropriate timing in a region where the load of the engine 1 is relatively high, and the SI rate is maintained at a predetermined SI rate. Is possible. Further, the amount of fresh air in the combustion chamber 17 can be increased by scavenging the burned gas.

  In the first intermediate load region (B1), the injector 6 is divided into the combustion chamber 17 in two stages of the pre-stage injection and the post-stage injection during the compression stroke, similarly to the second medium load area (B2). Inject fuel. The front-stage injection injects fuel at a timing away from the ignition timing, and the rear-stage injection injects fuel at a timing close to the ignition timing. The pre-stage injection may be performed, for example, during the first half of the compression stroke from the intake stroke, and the post-stage injection may be performed, for example, within the first half of the expansion stroke from the second half of the compression stroke.

  When the injector 6 performs the pre-stage injection and the post-stage injection within these periods, the entire combustion chamber 17 has an excess air ratio λ of 1.0 ± 0.2 and a G / F of 18 to A substantially homogeneous air-fuel mixture having a value of 50 is formed. Since the air-fuel mixture is substantially homogeneous, it is possible to improve fuel efficiency by reducing unburned loss and to improve exhaust gas performance by avoiding the generation of smoke. The excess air ratio λ is preferably 1.0 to 1.2.

  When the spark plug 25 ignites the air-fuel mixture at a predetermined timing before the compression top dead center, the air-fuel mixture burns by flame propagation. After the start of combustion by flame propagation, the unburned mixture self-ignites and CI burns. The fuel injected by the latter-stage injection mainly undergoes SI combustion. The fuel injected by the pre-stage injection mainly undergoes CI combustion. Since the pre-injection is performed during the compression stroke, it is possible to prevent the fuel injected by the pre-injection from inducing abnormal combustion such as premature ignition. Moreover, the fuel injected by the latter stage injection can be stably burned by flame propagation.

(High load range (high load SI combustion))
When the load on the engine 1 further increases and the operating state of the engine 1 enters the high load region (C), the engine 1 performs high load SI combustion. Therefore, the SI rate becomes 100% in the high load region (C).

  The throttle valve 43 is fully open. The supercharger 44 supercharges fresh air and external EGR gas even in the high load region (C). By changing the opening degree, the EGR valve 54 gradually reduces the amount of external EGR gas introduced as the load on the engine 1 increases. By doing so, the fresh air introduced into the combustion chamber 17 increases as the load on the engine 1 increases. As the amount of fresh air increases, the amount of fuel can be increased, which is advantageous in increasing the maximum output of the engine 1. The amount of EGR gas introduced into the combustion chamber 17 is continuous between the first medium load region (B1) and the high load region (C).

  Also in the high load region (C), similarly to the first intermediate load region (B1), an overlap period is provided in which both the intake valve 21 and the exhaust valve 22 are opened across the exhaust top dead center. The burned gas remaining in the combustion chamber 17 is scavenged by the supercharging pressure. Thereby, generation | occurrence | production of abnormal combustion is suppressed. In addition, the amount of fresh air in the combustion chamber 17 can be increased.

  In the region on the low rotation side of the high load region (C) (that is, the first high load region (C1)), as described above, the injector 6 injects fuel into the combustion chamber 17 within the retard period. . In the region on the high rotation side of the high load region (C) (that is, the second high load region (C2)), the injector 6 injects fuel into the combustion chamber 17 during the intake stroke. In any case, a substantially homogeneous air-fuel mixture is formed in the combustion chamber 17 with an excess air ratio λ of 1.0 ± 0.2 and a G / F of 18-50.

  At the maximum load, the excess air ratio λ is, for example, 0.8. Further, the G / F of the air-fuel mixture may be 17, for example, at the maximum load. When the spark plug 25 ignites the air-fuel mixture at a predetermined timing before the compression top dead center, the air-fuel mixture burns by flame propagation. In the high load region (C), the air-fuel mixture undergoes SI combustion without leading to self-ignition by high-pressure retarded injection or fuel injection during the intake stroke.

<Operation control of the engine in the rotational direction>
(Adjustment of SI rate)
FIG. 11 shows the relationship between the engine speed of the engine 1 and the SI rate at a predetermined load in the middle load region (B) where SPCCI combustion is performed. Like the driving | running | working area | region map 700 shown to FIG. 7A, the rotation speed N1 is the minimum rotation speed of a medium load area | region (B). The rotational speed N2 is the highest rotational speed in the medium load region (B) and is located at the boundary with the high rotational region (D) where SI combustion is performed (the same applies to the following).

  CI combustion has a disadvantage that the combustion noise is larger than SI combustion. However, generally, when the rotational speed of the engine 1 is low, the NVH of the engine 1 is small. For this reason, at the time of low rotation, even if the combustion noise of CI combustion is added to NVH, NVH can be maintained in a state below its allowable value. Therefore, when the rotational speed of the engine 1 is low, the ECU 10 can sufficiently perform the CI combustion by reducing the SI rate. Thereby, at the time of low rotation of the engine 1, NVH does not become a problem, and improvement in fuel consumption can be achieved.

  On the other hand, when the rotational speed of the engine 1 increases, the NVH of the engine 1 increases accordingly. When the combustion noise of CI combustion is added to the increased NVH, the NVH may exceed its allowable value. Therefore, the ECU 10 adjusts so that the SI rate increases as the rotational speed of the engine 1 increases.

  As shown in FIG. 11, the ECU 10 of this configuration example increases the SI rate linearly as the rotational speed of the engine 1 increases. Since the maximum rotational speed N2 is a boundary with the high rotational speed region (D) in which only SI combustion is performed, the SI rate at the rotational speed is adjusted to 100%. Therefore, it is possible to smoothly shift the combustion state between the medium load region (B) where the SPCCI combustion is performed and the high rotation region (D) where the SI combustion is performed without greatly fluctuating.

  That is, the CI combustion can be continuously reduced by adjusting the SI rate performed for the rotational speed of the engine 1, and combustion by self-ignition is performed in the high rotational direction to the limit. It becomes possible. As a result, in this engine 1, not only in the high load direction but also in the high rotation direction, the operation region in which combustion by self-ignition is performed can be expanded in a state where NVH, which may be a problem, is maintained below the allowable value. Therefore, this engine 1 is excellent in fuel consumption performance. Such adjustment of the SI rate can be realized by advanced operation control of the engine 1 executed by the ECU 10 (details will be described later).

  When the load on the engine 1 is high, the amount of fuel increases and the amount of heat received in the combustion chamber 17 also increases, so the temperature of the combustion chamber 17 becomes relatively high. Therefore, when the load is high, SI combustion in SPCCI combustion becomes faster than when the load is low. Therefore, when compared at the same rotation speed, the SI rate is adjusted to be higher when the load of the engine 1 is high than when the load of the engine 1 is low. Therefore, as indicated by a one-dot chain line in FIG. 11, the slope of the straight line indicating the change in the SI rate with respect to the engine speed (SI rate / engine speed) becomes gentler as the load on the engine 1 increases.

(Means to increase SI rate at high rotation)
At high speed, the combustion cycle is shorter than at low speed. Along with this, the amount of heat received by the inner wall of the combustion chamber during combustion is reduced, so that the temperature of the combustion chamber 17 is lower than that during low rotation. The air-fuel mixture is also less likely to receive heat from the combustion chamber 17. Furthermore, the time for which the air-fuel mixture burns is also shortened. At the time of high rotation, it is necessary to adjust so that the SI rate increases under such disadvantageous conditions.

  On the other hand, if the temperature of the air-fuel mixture itself in the combustion chamber 17 is increased, combustion can be promoted even under conditions where it is difficult to receive heat from the combustion chamber 17.

  CI combustion is ignited after SI combustion. When the temperature of the air-fuel mixture is increased, both SI combustion and CI combustion are promoted, but the combustion time is short at high revolutions. For this reason, the SI combustion in which the combustion starts first becomes faster and the combustion is promoted than the CI combustion in which the combustion starts later. Therefore, if the temperature of the air-fuel mixture is increased, the SI rate can be increased at the time of high rotation.

(First means)
Therefore, in the engine 1, the internal EGR rate is adjusted in the medium load region (B) where the SPCCI combustion is performed in order to change the SI rate with respect to the change in the rotational speed of the engine 1.

  Specifically, the ECU 10 outputs a control signal to the intake electric S-VT 23 and / or the exhaust electric S-VT 24 to change the opening timing of the intake valve 21 and / or the exhaust valve 22. By doing so, while providing a negative overlap period, the length is changed and the high-temperature burned gas (internal EGR gas) introduced into the combustion chamber 17 is set to a predetermined amount. As a result, the internal EGR rate, specifically, the “mass ratio (%) of the internal EGR gas to the total gas in the combustion chamber 17” is changed.

  As shown in FIG. 7A, in the engine 1 adopting the operation region map 700, the medium load region (B) has a first medium load region (B1) located on the high load side and a first load located on the low load side. 2 and the middle load region (B2). In the first medium load region (B1), in order to expand the operation region in which SPCCI combustion is performed in the high load direction while corresponding to the high load, supercharging is performed, whereas in the second medium load region (B2), No supercharging.

  In the engine 1, since a positive overlap period is provided in the first medium load region (B1) where supercharging is performed, even if the internal EGR gas is introduced, scavenging is performed by the supercharging pressure. Therefore, since it is difficult to change the internal EGR rate in the first medium load region (B1), this first means is not adopted. In the second medium load region (B2), the ECU 10 outputs a control signal to the intake electric S-VT 23 or the like so that the internal EGR rate becomes higher at the time of high rotation that is higher than when the rotation speed of the engine 1 is low and low.

  FIG. 12 (a) shows the relationship (internal EGR rate / engine speed) between the engine speed of the engine 1 and the internal EGR rate at a predetermined load in the second medium load region (B2). The ECU 10 increases the internal EGR rate linearly as the engine speed increases from the minimum engine speed N1. As a result, the temperature of the air-fuel mixture in the combustion chamber 17 increases, so that even if the amount of heat received from the combustion chamber 17 is small and the combustion time is shortened, SI combustion becomes rapid and the SI rate increases. When the SI rate increases, CI combustion decreases, so that combustion noise generated by SPCCI combustion can be suppressed. When the rotational speed of the engine 1 is high, NVH can be suppressed to an allowable value or less. The relationship between the engine speed and the internal EGR rate shown in FIG. 12 is an example, and the rate of change can be changed as appropriate according to the specifications.

  As shown in FIG. 12A, when the rotational speed of the engine 1 exceeds a predetermined rotational speed (restriction start rotational speed Na), the increase in the internal EGR rate is restricted. Specifically, a predetermined map is stored in the memory 102 of the ECU 10 in order to change the internal EGR rate in response to a change in the rotational speed of the engine 1, and the limit start rotational speed Na is included in the map. Is preset. The ECU 10 outputs a control signal to the intake electric S-VT 23 or the like so as to limit the increase in the internal EGR rate when the rotation speed of the engine 1 exceeds the limit start rotation speed Na.

  The limit start rotational speed Na is set to a value larger than at least the rotational speed obtained by dividing the engine rotational speed range (the rotational speed N1 to the rotational speed N2) in the medium load region (B) into two equal parts. In particular, it is preferable that the range of the engine speed in the medium load region (B) is set to a range on the high speed side when it is divided into three equal parts.

  In this configuration example, the internal EGR rate reaches its peak and is held at a substantially constant value by exceeding the limit start rotational speed Na (the value may slightly increase or decrease depending on the specification). obtain). If the internal EGR rate is increased and the temperature of the air-fuel mixture becomes excessively high, CI combustion also becomes rapid and combustion noise increases, which may exceed the allowable value of NVH. Therefore, the ECU 10 controls the NVH to be equal to or less than the allowable value by setting the limit start rotation speed Na before the rotation speed that can cause such a situation.

  When the load on the engine 1 is high, the amount of fuel increases and the amount of heat received in the combustion chamber 17 also increases, so the temperature of the combustion chamber 17 becomes relatively high. Therefore, when the load is high, the temperature of the air-fuel mixture becomes higher than when the load is low, and SI combustion in SPCCI combustion becomes rapid. Therefore, the ECU 10 changes the internal EGR rate so that the internal EGR rate decreases as the load on the engine 1 increases.

  At this time, the limit start rotational speed Na is set so as to be displaced toward the low speed side as the engine load increases. As the engine load increases, the temperature of the air-fuel mixture tends to become excessively high. Therefore, the higher the engine load is, the more reliably the NVH exceeds its allowable value by setting the limit start speed Na to the low speed side.

  FIG. 12A shows the relationship between the engine speed and the internal EGR rate when the internal EGR gas is introduced by setting the negative overlap period (also referred to as NVO setting). FIG. 12B shows the relationship between the engine speed and the internal EGR rate when the internal EGR gas is introduced by setting the positive overlap period (also referred to as PVO setting).

  In the introduction of the internal EGR gas by the PVO setting, unlike the introduction of the internal EGR gas by the NVO setting in which the high-temperature burned gas is confined in the combustion chamber 17, the high-temperature burned gas once flows out to the intake port 18. Then, it is again introduced into the combustion chamber 17. By flowing out to the intake port 18 having a low temperature, the high-temperature burned gas introduced into the combustion chamber 17, that is, the internal EGR gas is cooled.

  Therefore, in comparison with the introduction of the internal EGR gas with the NVO setting, the temperature of the combustion chamber 17 is relatively lowered with the introduction of the internal EGR gas with the PVO setting. Therefore, even if the internal EGR rate increases, the temperature of the air-fuel mixture is unlikely to become excessively high. Therefore, even if the setting of the limit start rotation speed Na is unnecessary or the setting of the limit start rotation speed Na is necessary, the value is set to a value sufficiently larger than the NVO setting.

  Therefore, in the introduction of the internal EGR gas by the PVO setting, the internal EGR rate continuously increases without being limited until reaching the vicinity of the rotation speed N2 that is the maximum rotation speed. At the time of high load shown in FIG. 12B, the limit start speed Na is set near the maximum speed N2, but it may not be necessary to set the limit start speed Na even at high load. obtain. Further, it may be necessary to set the limit start rotational speed Na even when the load is low or medium. However, even in that case, the limit start rotational speed Na is set to a value closer to the maximum rotational speed N2 as compared with the introduction of the internal EGR gas by the NVO setting.

  In the engine 1 that employs an operation region map 701 or an operation region map 702 to be described later, a swirl flow having a strength higher than a predetermined level is formed in the combustion chamber 17. When such a swirl ratio is formed in the combustion chamber 17, the mixture in the combustion chamber 17 can be stratified by the flow. As a result, the air-fuel mixture around the spark plug 25 can be brought into a state suitable for SI combustion, so that stable SPCCI combustion can be realized.

  Therefore, in this case, in order to ensure a high SI rate, even if the internal EGR rate is increased during high rotation, the EGR gas around the spark plug 25 can be reduced by stratification of the air-fuel mixture. Thereby, the stability of the ignition of SI combustion can be ensured, and stable SPCCI combustion can be realized.

(Second means)
With only the first means, only the second medium load region (B2) of the medium load region (B) performing SPCCI combustion can expand the operation region in which combustion by self-ignition is performed in the high rotation direction. Therefore, in the engine 1, in the first medium load region (B1), the temperature of the air-fuel mixture itself in the combustion chamber 17 is increased in the high rotation direction, similarly to the second medium load region (B2). A second means is also employed so that it can be changed.

  That is, the engine 1 changes the external EGR rate in the medium load region (B) in which SPCCI combustion is performed in order to change the SI rate with respect to a change in the rotational speed of the engine 1.

  Specifically, by changing the opening degree of the EGR valve 54, the amount of cooled burned gas (external EGR gas) cooled to a temperature lower than that of the internal EGR gas is changed into the combustion chamber 17, thereby The external EGR rate, specifically, “mass ratio (%) of external EGR gas to total gas in combustion chamber 17” is changed.

  The ECU 10 outputs a control signal to the EGR valve 54 so that the external EGR rate becomes lower during high rotation than during low rotation. The temperature of the external EGR gas is lower than that of the air-fuel mixture in the combustion chamber 17. For this reason, the greater the amount of external EGR gas introduced, the lower the temperature of the air-fuel mixture. Therefore, by lowering the external EGR rate, the temperature of the air-fuel mixture itself in the combustion chamber 17 can be changed higher.

  As shown in FIG. 7A, the external EGR gas is introduced in both the first medium load region (B1) and the second medium load region (B2), and particularly in the first medium load region (B1). Has been introduced. Therefore, the second means is effective in the first medium load region (B1), and in combination with the first means, the engine 1 is effectively effective in the entire region of the medium load region (B) where the SPCCI combustion is performed. The SI rate can be adjusted to be high with respect to the change in the rotation speed.

  FIG. 13A shows the relationship between the level of the engine 1 and the external EGR rate (external EGR rate / engine speed) at a predetermined load in the middle load region (B). The solid line indicates the relationship in the first medium load region (B1), and the one-dot broken line indicates the relationship in the second medium load region (B2).

  As described above, in the engine 1 that employs the operation region map 700, the swirl control valve 56 is controlled to be substantially fully opened in each operation region including the intermediate load region (B) in which SPCCI combustion is performed. FIG. 13A shows the above-described relationship in the case where the swirl flow is hardly generated in the combustion chamber 17 or the flow is weak even if the swirl flow is generated. .

  The ECU 10 keeps the external EGR rate substantially constant from the minimum rotational speed N1 to a predetermined rotational speed (decrease start rotational speed Nb). By doing so, the EGR rate (external EGR rate) can be set to a high value at the time of low rotation, so even if the G / F value of the air-fuel mixture that must be secured is high, it can be within an appropriate range. it can. Since the dilution ratio of the air-fuel mixture can be kept high, the fuel efficiency can be improved.

  Then, when the reduction start rotational speed Nb is exceeded, the ECU 10 reduces the external EGR rate according to the load of the engine 1. Specifically, a predetermined map is stored in the memory 102 of the ECU 10 in order to change the external EGR rate in response to a change in the rotational speed of the engine 1, and the reduction start rotational speed Nb is included in the map. Is preset.

  The reduction start rotational speed Nb is also set to a value that is at least larger than the rotational speed obtained by dividing the engine rotational speed range in the medium load region (B) into two equal parts, similarly to the limited start rotational speed Na. In particular, it is preferable to set the range of the engine speed in the medium load region (B) to the high speed side when the engine speed is divided into three equal parts, and the decrease start speed Nb and the limit start speed Na are set to the same speed. It is preferable to do this.

  Based on the map, the ECU 10 outputs a control signal to the EGR valve 54 so that the external EGR rate starts decreasing according to the load of the engine 1 when the decrease starting rotational speed Nb is exceeded. When the load is low, the value of the external EGR rate may be maintained as it is even if it exceeds the decrease start rotational speed Nb.

  As a result, in the high speed region where the reduction start rotational speed Nb or higher, the ratio of the external EGR gas introduced into the combustion chamber 17 decreases, and the temperature of the mixture itself increases. Therefore, even if the amount of heat received from the combustion chamber 17 is small and the combustion time is shortened, SI combustion becomes rapid and the SI rate increases. When the SI rate increases, CI combustion decreases, so that combustion noise generated by SPCCI combustion can be suppressed. When the rotational speed of the engine 1 is high, NVH can be suppressed to an allowable value or less. Note that the relationship between the engine speed and the external EGR rate shown in FIG. 13 is an example, and the rate of change can be appropriately changed according to the specifications.

  At this time, the ECU 10 outputs a control signal to the EGR valve 54 so that the external EGR rate suddenly decreases (the reduction rate increases) as the load on the engine 1 increases.

  As the load on the engine 1 increases, the amount of fuel increases accordingly. If the amount of burned gas is excessive with respect to the increasing fuel, ignition by spark ignition, that is, SI combustion in SPCCI combustion becomes unstable, and the SI rate may decrease. Therefore, as the load on the engine 1 increases, the rate of decrease in the external EGR rate is increased to stabilize ignition, thereby maintaining a high SI rate.

  The ECU 10 also changes the external EGR rate in the second medium load region (B2), similarly to the first medium load region (B1). In this configuration example, the EGR gas introduced in the first medium load region (B1) is only the external EGR gas, whereas the internal EGR gas is also introduced in the second medium load region (B2). Therefore, in the second intermediate load region (B2), the external EGR rate is changed in consideration of the state quantity in the combustion chamber 17 such as the G / F of the air-fuel mixture in relation to the introduction amount of the internal EGR gas. The As a result, the external EGR rate in the second medium load region (B2) is relatively low and has a different value compared to the first medium load region (B1).

  In the engine 1 that employs an operation region map 701 or an operation region map 702 described later, a swirl flow having a strength greater than or equal to a predetermined level is formed in the combustion chamber 17, unlike the engine 1 that employs the operation region map 700. . FIG. 13B shows the above-described relationship in such an engine 1.

  When a swirl flow having a predetermined strength or more is formed in the combustion chamber 17, the fluidity of the air-fuel mixture increases in the combustion chamber 17. Therefore, even if the amount of EGR gas is large, the amount of EGR gas around the spark plug 25 can be reduced when ignition is performed. Therefore, even if the load increases and the fuel increases, the stability of ignition by ignition can be ensured, so that SI combustion in SPCCI combustion can be performed stably.

  Thereby, as shown in FIG. 13B, the ECU 10 controls the EGR valve 54 so that the rate of decrease in the external EGR rate is maintained substantially constant with respect to the change in the load of the engine 1. Is output. The ECU 10 reduces the external EGR rate in the same state as when the load is low even when the load is high.

(Third means)
In the method of increasing the temperature of the air-fuel mixture as in the first and second means, if the temperature becomes excessively high, the CI combustion also becomes rapid and the combustion noise increases and exceeds the allowable value of NVH. There is a fear. In addition, since the internal EGR rate and the external EGR rate are limited by other state quantities such as the G / F of the air-fuel mixture, the effect of suppressing the combustion noise in the SPCCI combustion is limited only by changing the internal EGR rate and the external EGR rate. May not be obtained.

  In particular, in the first means, the increase in the internal EGR rate is limited within a range exceeding the limit start rotational speed Na. Also in the second means, if the external EGR rate is excessively decreased within the range where the decrease start rotational speed Nb is exceeded, the temperature of the air-fuel mixture becomes excessively high and combustion noise increases, exceeding the allowable value of NVH. The rate of decline is limited. Accordingly, there is a case where the desired SI rate is not increased in a range exceeding the limit start rotation speed Na and the decrease start rotation speed Nb. In that case, the effect of suppressing combustion noise is limited, and a sufficient effect cannot be obtained.

  Therefore, in this engine 1, another means different from the method of increasing the temperature of the air-fuel mixture is used, and the SI rate is also increased by this.

  That is, the ECU 10 uses the third means for strengthening the intake air flow by controlling the swirl control valve 56 in combination with the first and second means. When the intake air flow is strengthened, the SI combustion becomes rapid and the SI rate increases. As a result, the combustion noise of SPCCI combustion can be suppressed.

  FIG. 14A shows the relationship between the rotational speed of the engine 1 adopting the operation region map 700 and the opening degree of the swirl control valve 56. As described above, in the engine 1 that employs the operation region map 700, the swirl control valve 56 is controlled to be substantially fully opened in each operation region including the intermediate load region (B) in which SPCCI combustion is performed. FIG. 14A shows the above-described relationship in that case. On the other hand, FIG. 14B shows the above-described relationship in the engine 1 adopting the operation region map 701 and the engine 1 adopting the operation region map 702, which will be described later.

  As shown in FIG. 14 (a), in the engine 1 that employs the operation region map 700, the rotational speed of the engine 1 is such that the restriction start rotational speed Na at which the restriction of the internal EGR rate starts or the external EGR rate decreases. When the ECU 10 reaches a predetermined rotation speed (operation start rotation speed Nc) that is difficult to adjust to the desired SI rate by only the first or second means, such as the decrease start rotation speed Nb at which the ECU 10 starts, the ECU 10 A control signal is output to the control valve 56, and the opening degree is changed from substantially fully open to closed side. Thereby, the swirl flow in the combustion chamber 17 is strengthened. The ECU 10 linearly changes the opening of the swirl control valve 56 to the closed side as the rotational speed of the engine 1 increases. As the rotation speed of the engine 1 increases, the swirl flow becomes stronger, and therefore SI combustion becomes more rapid.

  As a result, in the first and second means, it is possible to suppress the combustion noise of SPCCI combustion even in the range on the high speed side that is restricted. As a result, when the rotational speed of the engine 1 is high, NVH can be suppressed to an allowable value or less. The combination of the first and second means and the third means makes it possible to change the SI rate higher and more appropriately.

  On the other hand, as shown in FIG. 14B, the engine 1 employing the operation region map 701 and the engine 1 employing the operation region map 702, which will be described later, have a predetermined strength in the region where the SPCCI combustion is performed. A swirl flow is formed in the combustion chamber 17.

  For example, at the timing from intake to ignition, the swirl control valve 56 is controlled at a predetermined opening on the closing side without being fully closed so that the swirl ratio is in the range of 1.5 to 3. The opening degree or swirl ratio of the swirl control valve 56 is appropriately set according to the specifications of the engine 1. FIG. 14B illustrates a case where the swirl control valve 56 is held substantially constant at an opening degree of about 30% with respect to a change in the rotational speed of the engine 1.

  By forming such a swirl flow in the combustion chamber 17 at the time of performing SPCCI combustion, the fluidity in the combustion chamber 17 can be improved, and the stratified mixing in the combustion chamber 17 It becomes possible to form a mind. Thereby, stable SPCCI combustion can be realized in a wide operation region of the engine 1. Details thereof will be described later.

(Fourth means)
The engine 1 that employs the operation region map 700 is configured such that the SI rate is increased toward the high rotation direction even when the ignition timing is changed.

  FIG. 15 shows the relationship between the engine speed of the engine 1 and the ignition timing at a predetermined load in the middle load region (B) where SPCCI combustion is performed. Similar to the third means, the engine 1 has a rotational speed such as a limiting start rotational speed Na at which the restriction of the internal EGR rate is started and a reduction starting rotational speed Nb at which the reduction of the external EGR rate is started. When the ECU 10 reaches a predetermined rotation speed (advance start rotation speed Nd) that is difficult to adjust to the desired SI rate with only the second means, the ECU 10 outputs a control signal to the spark plug 25 to advance the ignition timing. Horn. Thereby, the start of SI combustion is accelerated and SI combustion becomes rapid. The ECU 10 linearly advances the ignition timing as the rotational speed of the engine 1 increases. As the rotational speed of the engine 1 increases, the SI combustion becomes more rapid.

  The fourth means can be used in place of the third means. The fourth means may be used together with the third means. By appropriately combining the first to fourth means, it is possible to stably adjust the SI rate even at high revolutions.

(Change of SI rate)
FIG. 16 shows a flow relating to operation control of the engine 1 executed by the ECU 10. The ECU 10 determines the operating state of the engine 1 based on the detection signals of the sensors SW1 to SW16, and the combustion chamber 17 so that the combustion in the combustion chamber 17 becomes combustion at the SI rate corresponding to the operating state. The change of the state quantity, the change of the injection quantity, the change of the injection timing, and the change of the ignition timing are performed.

  The ECU 10 also changes the SI rate by the SI rate changing means 102a when it is determined that the SI rate needs to be changed based on the detection signal of each sensor.

  First, in step S1, the ECU 10 reads the detection signals of the sensors SW1 to SW16. Next, in step S2, the ECU 10 determines the operating state of the engine 1 based on the detection signal and sets a target SI rate. The target SI rate is as shown in FIG. 9 or FIG.

  In the subsequent step S3, the ECU 10 sets a target in-cylinder state quantity for realizing the set target SI rate based on a preset combustion model. Specifically, the target temperature and target pressure in the combustion chamber 17 and the target state quantity are set. In step S4, 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 intake electric motor S-VT 23 and the exhaust electric motor necessary for realizing the target in-cylinder state quantity. The phase angle of the S-VT 24 and the opening degree of the swirl control valve 56 are set.

  The ECU 10 sets control amounts of these devices based on a map that is set in advance and stored in the ECU 10. The ECU 10 outputs control signals to the EGR valve 54, the throttle valve 43, the air bypass valve 48, the intake electric S-VT 23 and the exhaust electric S-VT 24, and the swirl control valve 56 based on the set control amount. As each device operates based on the control signal of the ECU 10, the state quantity in the combustion chamber 17 becomes the target state quantity.

  The ECU 10 further calculates a predicted value and an estimated value of the state 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, and 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. As will be described later, the setting of the fuel injection amount in the compression stroke and the ignition timing are set. Used for setting. The state quantity estimated value is also used for calculation of a state quantity error by comparison with an actual combustion state, as will be described later.

  In step S5, the ECU 10 sets the fuel injection amount during the intake stroke based on the state amount predicted value. When fuel is not injected during the intake stroke, the fuel injection amount is zero. In step S6, the ECU 10 controls the injection of the injector 6. That is, a control signal is output to the injector 6 so that fuel is injected into the combustion chamber 17 at a predetermined injection timing.

  In step S7, 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. When performing divided injection during the compression stroke, the injection amount of the front-stage injection and the injection amount of the rear-stage injection are respectively set. In step S8, the ECU 10 outputs a control signal to the injector 6 so as to inject fuel into the combustion chamber 17 at an injection timing based on a preset map.

  In step S9, 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 S10, the ECU 10 outputs a control signal to the spark plug 25 so as to ignite 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 S11, the ECU 10 reads the pressure change in the combustion chamber 17 detected 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 S12, the ECU 10 also compares the detection result of the combustion state with the state quantity estimated value estimated in step S4, and calculates an error between the state quantity estimated value and the actual state quantity. The calculated error is used for estimation in step S4 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, the intake electric S-VT 23 and the exhaust electric S-VT 24 so that the state quantity error is eliminated. Change the phase angle. Thereby, the amount of fresh air and EGR gas introduced into the combustion chamber 17 is changed. This feedback of the state quantity error corresponds to changing the SI rate when the ECU 10 determines that the SI rate needs to be changed based on the error between the target SI rate and the actual SI rate.

  The ECU 10 also performs injection during the compression stroke so that the ignition timing can be advanced when it is predicted in step S8 that the temperature in the combustion chamber 17 is 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 it is predicted in step S7 that the temperature in the combustion chamber 17 will be higher than the target temperature based on the state quantity estimated value, 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, as shown in P2 of Fig. 17, when the low temperature of the combustion chamber 17, after the SI combustion by spark ignition is started, will be the timing theta _CI unburned air-fuel mixture is self-ignition delay, SI ratio However, it will deviate from the target SI rate (see P1). In this case, unburned fuel increases and exhaust gas performance decreases.

Therefore, when the temperature inside the combustion chamber 17 is expected to be lower than the target temperature, ECU 10 is adapted to advance the injection timing, in step S10 of FIG. 17, it advances the ignition timing theta _IG. As shown in P3 of FIG. 17, since the start of SI combustion is accelerated, it is possible to generate sufficient heat by SI combustion. Therefore, when the temperature in the combustion chamber 17 is low, the self-ignition of the unburned mixture is performed. It is possible to prevent the timing θ _{CI from} being delayed. As a result, the SI rate approaches the target SI rate. An increase in unburned fuel and a decrease in exhaust gas performance are prevented.

  Moreover, as shown in P4 of FIG. 17, when the temperature in the combustion chamber 17 is high, immediately after the SI combustion is started by spark ignition, the unburned mixture self-ignites, and the SI rate becomes the target. It deviates from the SI rate (see P1). In this case, combustion noise increases.

Therefore, when the temperature inside the combustion chamber 17 is expected to be higher than the target temperature, ECU 10 is adapted to retard the injection timing, in step S10 of FIG. 16, retarding the ignition timing theta _IG. As shown in P5 of FIG. 17, since the start of the SI combustion becomes slow, when the temperature inside the combustion chamber 17 is high, it is possible to prevent the timing theta _CI autoignition of the unburned air-fuel mixture becomes faster it can. As a result, the SI rate approaches the target SI rate. An increase in combustion noise is avoided.

  These changes in the injection timing and the ignition timing correspond to the change in the SI rate when the ECU 10 determines that the SI rate in the SPCCI combustion needs to be changed. By changing the injection timing, an appropriate air-fuel mixture can be formed in the combustion chamber 17 at the ignition timing that is advanced or retarded. The spark plug 25 can surely ignite the air-fuel mixture, and the unburned air-fuel mixture can self-ignite at an appropriate timing.

  In FIG. 17, the combustion chamber 17 is controlled through the control of the throttle valve 43, the EGR valve 54, the air bypass valve 48, the intake electric S-VT 23, the exhaust electric S-VT 24, and the swirl control valve 56 based on the actual combustion state. The point of changing the state quantity is as described in step S12 and step S4 of FIG.

  The engine 1 changes the SI rate by a state quantity setting device including a throttle valve 43, an EGR valve 54, an air bypass valve 48, an intake electric S-VT 23, an exhaust electric S-VT 24, and a swirl control valve 56. By changing the state quantity in the combustion chamber 17, the SI rate can be roughly changed. At the same time, the engine 1 changes the SI rate by changing the fuel injection timing and the ignition timing. By changing the injection timing and the ignition timing, for example, the difference between cylinders can be corrected, or the self-ignition timing can be finely changed. By changing the SI rate in two stages, the engine 1 can accurately realize the target SPCCI combustion corresponding to the operating state.

<Second operation area map>
FIG. 7B shows a second configuration example (second operation region map 701) of the operation region map of the engine 1. The driving area map 701 is divided into the following three areas.

(A) Low load region including idle operation (B) Medium load region between low load region (A) and next high load region (C) (C) High load region including fully open load

  In other words, the operation region map 701 has a low load region (A), a medium load region (B), and a high load region (C) in the direction of the rotational speed of the engine 1 as compared with the operation region map 700 described above. Each of them has been enlarged.

  In the engine 1 employing the operation region map 701, in particular, a swirl flow having a predetermined strength is generated in the combustion chamber 17 in order to expand the medium load region (B) in which the SPCCI combustion is performed. ing.

  For example, the ECU 10 adjusts the opening degree of the swirl control valve 56 as necessary. When the operating state of the engine 1 is in the medium load region (B), the swirl control valve 56 is fully closed or has a predetermined opening on the closing side. As a result, a swirl flow having a predetermined strength is formed in the combustion chamber 17. At the timing from intake to ignition, the swirl ratio may be 4 or more. Moreover, a swirl ratio is good also as less than 4, for example, the range of 1.5-3.

  Also in the engine 1 that employs the operation region map 701, it is preferable to adjust the SI rate similarly to the engine 1 that employs the operation region map 700. That is, as shown in FIG. 11, in the second medium load region (B2) in which supercharging is not performed, the SI rate is increased linearly as the rotational speed of the engine 1 increases. By doing so, SPCCI combustion can be stably performed in a state where NVH, which may be a problem, is maintained below an allowable value even in a high rotation region.

  Also, in the engine 1 that employs the operation region map 701, similarly to the engine 1 that employs the operation region map 700, the SI rate can be increased at the time of high rotation by the first to fourth means.

<Third driving area map>
FIG. 7C shows a third configuration example (third operation region map 702) of the operation region map of the engine 1. The driving area map 702 is divided into the following five areas.

(1) -1: Low load region including idle operation and extending to low and medium rotation regions (1) -2: Medium load higher than low load region and extending to low and medium rotation regions Load region (2): Medium load region where load is higher than medium load region and high load region including fully open load (3): Low rotation region where rotation speed is lower than medium rotation region in high load region ( 4): A low load region, a medium load region, a high load medium rotation region, and a high rotation region having a higher rotational speed than the high load low rotation region.

  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 middle rotation region, and the high rotation region may be used. In the operation region map 702, 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. The rotation speed N1 and the rotation speed N2 may be the same or different in each of the operation region map 700 and the operation region map 702.

  In the driving region map 702, the main purpose is to improve fuel consumption and exhaust gas performance, in the low load region (1) -1, the medium load region (1) -2, and the high load medium rotation region (2). The engine 1 performs SPCCI combustion. It differs from the operation region map 701 in that SPCCI combustion is performed when the engine 1 is operated at a low load and when the engine 1 is operated at a high load. The engine 1 also performs combustion by spark ignition in other regions, specifically, in the high load low rotation region (3) and the high rotation region (4). Hereinafter, the operation of the engine 1 in each region will be described in detail with reference to the fuel injection timing and the ignition timing shown in FIG.

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

  Reference numeral 601 in FIG. 18 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 indicated by reference numeral 601 in the low load region (1) -1. In addition, an example of each combustion waveform (that is, a waveform indicating a change in heat generation rate with respect to the crank angle, reference numeral 6014) is shown.

  When the engine 1 is operating in the low load region (1) -1, a swirl flow having a predetermined strength is formed in the combustion chamber 17 (see FIG. 14B). The swirl ratio when the engine 1 operates in the low load region (1) -1 is approximately 2, and is adjusted in a range of less than 4, for example, in a range of 1.5 to 3. Note that a strong swirl flow may be formed by setting the swirl ratio to 4 or more.

  The swirl flow is strong at the outer periphery of the combustion chamber 17 and weak at the center. The swirl control valve (SCV) 56 is controlled at a predetermined opening on the closing side without being fully closed. 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.

  When the engine 1 operates in the low load region (1) -1, 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. In other words, the excess air ratio λ of the air-fuel mixture exceeds 1 in the entire combustion chamber 17. 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.

  The EGR system 55 introduces EGR gas into the combustion chamber 17 as necessary when the engine 1 is operating in the low load region (1) -1. For example, the internal EGR gas is introduced by setting a positive overlap period.

  When the engine 1 operates in the low load region (1) -1, the air-fuel mixture is stratified between the central portion and the outer peripheral portion in the combustion chamber 17. A central portion in the combustion chamber 17 is a portion where the spark plug 25 is disposed, and an outer peripheral portion is a portion around the central portion and in contact with the liner of the cylinder 11. You may define the center part in the combustion chamber 17 as a part with a weak swirl flow, and an outer peripheral part as a part with a strong swirl flow.

  The fuel concentration of the air-fuel mixture at the center is higher than the fuel concentration at the outer periphery. 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.

  When the engine 1 operates in the low load region (1) -1, the injector 6 basically injects the fuel into the combustion chamber 17 in a plurality of times during the compression stroke. The air-fuel mixture is stratified in the central portion and the outer peripheral portion of the combustion chamber 17 by the divided fuel injection and the swirl flow in the combustion chamber 17.

  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 (see reference numeral 6013). Since the air-fuel mixture in the central portion 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. As a result, when the engine 1 is operated in the low load region (1) -1, both the suppression of the generation of combustion noise and the improvement of the fuel consumption performance due to the shortening of the combustion period are compatible.

(Medium load area (1) -2)
Even when the engine 1 is operating in the medium load region (1) -2, the engine 1 performs SPCCI combustion as in the low load region (1) -1. The intermediate load area (1) -2 corresponds to the intermediate load area (B) in the operation area map 701.

  Reference numeral 602 in FIG. 18 indicates fuel injection timing (reference numerals 6021 and 6022) and ignition timing (reference numeral 6023) when the engine 1 is operating in the operation state indicated by reference numeral 602 in the middle load region (1) -2. An example of each combustion waveform (reference numeral 6024) is shown.

  The EGR system 55 introduces EGR gas into the combustion chamber 17 when the operating state of the engine 1 is in the medium load region (1) -2. For example, the internal EGR gas is introduced by setting a positive overlap period.

  Even when the engine 1 operates in the medium load region (1) -2, the combustion chamber 17 has a predetermined strength with a swirl ratio of about 2, as in the low load region (1) -1. A swirl flow is formed. The swirl control valve (SCV) 56 is not fully closed but has a predetermined opening on the closing side. By forming the swirl flow, the residual gas accumulated in the cavity 31 can be driven out of the cavity 31. As a result, the G / F of the air-fuel mixture in the SI portion near the spark plug 25 and the G / F of the air-fuel mixture in the CI portion around the SI portion can be made different. As a result, as described above, if the total G / F of the entire combustion chamber 17 is set to 18 or more and 50 or less, SPCCI combustion can be stabilized.

  Further, since the turbulent energy in the combustion chamber 17 is increased by forming the swirl flow, when the engine 1 is operated in the medium load region (1) -2, the SI combustion flame is quickly propagated. SI combustion is stabilized. The controllability of CI combustion is enhanced by the stabilization of SI combustion. By optimizing the timing of CI combustion in SPCCI combustion, it is possible to suppress the generation of combustion noise and improve fuel efficiency. Further, torque variation between cycles can be suppressed.

  When the engine 1 is operated in the medium load region (1) -2, 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 catalyst purifies 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. Therefore, the excess air ratio λ of the air-fuel mixture may be set to 1.0 ± 0.2.

  When the engine 1 operates in the medium load region (1) -2, the injector 6 performs fuel injection during the intake stroke (reference numeral 6021) and fuel injection during the compression stroke (reference numeral 6022). By performing the first injection 6021 during the intake stroke, the fuel can be distributed substantially uniformly in the combustion chamber 17. By performing the second injection 6022 during the compression stroke, the temperature in the combustion chamber 17 can be lowered by the latent heat of vaporization of the fuel. It is possible to prevent the air-fuel mixture including the fuel injected by the first injection 6021 from being prematurely ignited.

  When the injector 6 performs the first injection 6021 during the intake stroke and the second injection 6022 during the compression stroke, the excess air ratio λ is 1.0 ± 0.2 in the combustion chamber 17 as a whole. An air-fuel mixture is formed. Since the fuel concentration of the air-fuel mixture is substantially uniform, it is possible to improve fuel efficiency by reducing unburned loss and improve exhaust gas performance by avoiding the generation of smoke. The excess air ratio λ is preferably 1.0 to 1.2. The total G / F of the entire combustion chamber 17 is 18 or more and 50 or less, and the G / F of the SI portion in the vicinity of the spark plug 25 is 14-22.

  When the spark plug 25 ignites the air-fuel mixture at a predetermined timing before the compression top dead center (reference numeral 6023), the air-fuel mixture burns by flame propagation. After the start of combustion by flame propagation, the unburned mixture self-ignites at the target timing and performs CI combustion. The fuel injected by the latter-stage injection mainly undergoes SI combustion. The fuel injected by the pre-stage injection mainly undergoes CI combustion. By setting the total G / F of the entire combustion chamber 17 to 18 to 50 and the G / F of the SI portion in the vicinity of the spark plug 25 to 14 to 22, SPCCI combustion can be stabilized.

  Here, as shown to FIG. 7C, the area | region (refer S / C OFF) where the supercharger 44 is turned off is a part of low load area | region (1) -1, and medium load area | region (1)- Part of 2. Specifically, the supercharger 44 is turned off in the low rotation side region in the low load region (1) -1. In the region on the high rotation side in the low load region (1) -1, the supercharger 44 is turned on in order to ensure a necessary intake charge amount corresponding to the increase in the rotational speed of the engine 1. Increase the boost pressure. Further, the supercharger 44 is turned off in the low load / low rotation side region in the medium load region (1) -2, and the fuel injection amount in the high load side region in the medium load region (1) -2. The turbocharger 44 is turned on in order to ensure the necessary intake charge amount corresponding to the increase in the engine speed, and is necessary in response to the increase in the rotational speed of the engine 1 in the high speed region. In order to ensure the intake charge amount, the supercharger 44 is turned on.

  Note that the supercharger 44 is turned on throughout the high load medium rotation region (2), the high load low rotation region (3), and the high rotation region (4).

(High load medium rotation range (2))
Even when the engine 1 is operating in the high load mid-rotation region (2), the engine 1 performs SPCCI combustion in the same manner as the low load region (1) -1 and the medium load region (1) -2.

  Reference numeral 603 in FIG. 18 denotes a fuel injection timing (reference numerals 6031 and 6032) and an ignition timing (reference numeral 6033) when the engine 1 is operating in the operating state of the reference numeral 603 in the high load mid-rotation region (2). An example of each combustion waveform (reference numeral 6034) is shown. Further, reference numeral 604 in FIG. 18 shows an example of each of the fuel injection timing (reference numeral 6041) and the ignition timing (reference numeral 6042) and the combustion waveform (reference numeral 6043) when the rotational speed is higher than the operating state of reference numeral 603. ing.

  The EGR system 55 introduces EGR gas (for example, external EGR gas) into the combustion chamber 17 when the operating state of the engine 1 is in the high load mid-rotation region (2). The engine 1 reduces the amount of EGR gas as the load increases. At full load, EGR gas may be zero.

  Further, when the engine 1 is operated in the high load mid-rotation region (2), the combustion chamber 17 has a predetermined strong force with a swirl ratio of less than 4, as in the low load region (1) -1. A swirl flow having a thickness is formed. A strong swirl flow may be formed with a swirl ratio of 4 or more. The swirl control valve (SCV) 56 is controlled at a predetermined opening on the closing side without being fully closed.

  When the engine 1 operates in the high load mid-rotation region (2), the air-fuel ratio (A / F) of the air-fuel mixture is richer than the stoichiometric air-fuel ratio or the stoichiometric air-fuel ratio in the entire combustion chamber 17 (that is, The excess air ratio λ of the mixture is λ ≦ 1).

  When the engine 1 operates on the low rotation side in the high load mid-rotation region (2), the injector 6 injects fuel in the intake stroke (reference numeral 6031) and injects fuel at the end of the compression stroke (reference numeral 6032). ). The end of the compression stroke may be the end when the compression stroke is divided into three equal parts in the initial, middle and final stages.

  The pre-injection 6031 that starts in the intake stroke may start fuel injection in the first 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 equal parts. Specifically, the pre-injection may start fuel injection at 280 ° CA before top dead center.

  When the injection start of the front injection 6031 is in the first half of the intake stroke, although not shown in the drawing, a part of the fuel enters the squish area 171 of the combustion chamber 17 by the fuel spray hitting the opening edge of the cavity 31, The remaining fuel enters the area within the cavity 31. The swirl flow is strong at the outer periphery of the combustion chamber 17 and weak at the center. Therefore, a part of the fuel that has entered the squish area 171 enters the swirl flow, and the remaining fuel that has entered the region within the cavity 31 enters the inside of the swirl flow. The fuel that has entered the swirl flow remains in the swirl flow during the intake stroke to the compression stroke, and forms an air-fuel mixture for CI combustion at the outer peripheral portion of the combustion chamber 17. The fuel that has entered the swirl flow also remains inside the swirl flow during the intake stroke to the compression stroke, and forms an air-fuel mixture for SI combustion in the central portion of the combustion chamber 17.

  When the engine 1 operates in the high load mid-rotation region (2), the air-fuel mixture at the center where the spark plug 25 is disposed preferably has an excess air ratio λ of 1 or less, and the air-fuel mixture at the outer periphery is The excess air ratio λ is 1 or less, preferably less than 1. The air-fuel ratio (A / F) of the air-fuel mixture in the center may be, for example, 13 or more and the theoretical air-fuel ratio (14.7) or less. The air-fuel ratio of the air-fuel mixture at the center may be leaner than the stoichiometric air-fuel ratio. The air-fuel ratio of the air-fuel mixture at the outer peripheral portion may be, for example, 11 or more and the stoichiometric air-fuel ratio or less, preferably 11 or more and 12 or less. When the excess air ratio λ at the outer peripheral portion of the combustion chamber 17 is less than 1, the amount of fuel in the air-fuel mixture increases at the outer peripheral portion, so the temperature can be lowered by the latent heat of vaporization of the fuel. The air-fuel ratio of the entire air-fuel mixture in the combustion chamber 17 may be 12.5 or more and the theoretical air-fuel ratio or less, preferably 12.5 or more and 13 or less.

  For example, the post-injection 6032 performed at the end of the compression stroke may start fuel injection at 10 ° CA before top dead center. By performing the post-stage injection immediately before the top dead center, the temperature in the combustion chamber can be lowered by the latent heat of vaporization of the fuel. The fuel injected by the pre-injection 6031 undergoes a low-temperature oxidation reaction during the compression stroke and shifts to a high-temperature oxidation reaction before the top dead center. However, the post-injection 6032 is performed immediately before the top dead center to burn the fuel. By lowering the indoor temperature, it is possible to suppress the transition from the low temperature oxidation reaction to the high temperature oxidation reaction, and it is possible to suppress the occurrence of premature ignition. In addition, the ratio of the injection quantity of front | former stage injection and the injection quantity of back | latter stage injection is good also as 95: 5 as an example.

  The spark plug 25 ignites the air-fuel mixture in the center of the combustion chamber 17 in the vicinity of the compression top dead center (reference numeral 6033). The spark plug 25 performs ignition after the compression top dead center, for example. Since the spark plug 25 is disposed at the center of the combustion chamber 17, the air-fuel mixture at the center starts SI combustion by flame propagation by the ignition of the spark plug 25.

  In the high load region, the fuel injection amount increases and the temperature of the combustion chamber 17 also increases, so that the CI combustion is likely to start early. In other words, pre-ignition of the air-fuel mixture tends to occur in the high load region. However, as described above, since the temperature of the outer peripheral portion of the combustion chamber 17 is lowered by the latent heat of vaporization of the fuel, it is avoided that the CI combustion starts immediately after the mixture is sparked. Can do.

  As described above, when the spark plug 25 ignites the air-fuel mixture in the central portion, the SI combustion is stabilized by increasing the combustion speed due to the high turbulent energy, and the flame of the SI combustion is stabilized in the combustion chamber 17. Propagating in the circumferential direction along the swirl flow. Thus, the unburned mixture self-ignites at a predetermined circumferential position in the outer peripheral portion of the combustion chamber 17, and CI combustion starts.

  In this SPCCI combustion concept, sufficient SI combustion is performed before the start of CI combustion by stratifying the air-fuel mixture in the combustion chamber 17 and generating a swirl flow in the combustion chamber 17. Can do. As a result, the generation of combustion noise can be suppressed, and the generation of NOx is also suppressed without the combustion temperature becoming too high. Further, torque variation between cycles can be suppressed.

  Moreover, since the temperature of the outer peripheral portion is low, CI combustion becomes moderate, and the generation of combustion noise can be suppressed. Furthermore, since the combustion period is shortened by CI combustion, torque can be improved and thermal efficiency can be improved in a high load region. Therefore, the engine 1 can improve fuel efficiency while avoiding combustion noise by performing SPCCI combustion in a high load region.

  When the engine 1 operates on the high rotation side in the high load mid-rotation region (2), the injector 6 starts fuel injection in the intake stroke (reference numeral 6041).

  The pre-stage injection 6041 that starts in the intake stroke may start fuel injection in the first half of the intake stroke, as described above. Specifically, the front injection 6041 may start fuel injection at 280 ° CA before top dead center. The end of the front injection may exceed the intake stroke and be in the compression stroke. By starting the injection of the front injection 6041 in the first half of the intake stroke, an air-fuel mixture for CI combustion is formed in the outer peripheral portion of the combustion chamber 17 and an air-fuel mixture for SI combustion is formed in the central portion of the combustion chamber 17 can do. As described above, the air-fuel mixture in the center portion where the spark plug 25 is disposed preferably has an excess air ratio λ of 1 or less, and the air-fuel mixture in the outer peripheral portion has an air excess ratio λ of 1 or less, preferably 1 Is less than. The air-fuel ratio (A / F) of the air-fuel mixture in the center may be, for example, 13 or more and the theoretical air-fuel ratio (14.7) or less. The air-fuel ratio of the air-fuel mixture at the center may be leaner than the stoichiometric air-fuel ratio. The air-fuel ratio of the air-fuel mixture at the outer peripheral portion may be, for example, 11 or more and the stoichiometric air-fuel ratio or less, preferably 11 or more and 12 or less. The air-fuel ratio of the entire air-fuel mixture in the combustion chamber 17 may be 12.5 or more and the theoretical air-fuel ratio or less, preferably 12.5 or more and 13 or less.

  When the rotation speed of the engine 1 increases, the time for the fuel injected by the pre-injection 6041 to react becomes shorter. Therefore, the latter stage injection for suppressing the oxidation reaction of the air-fuel mixture can be omitted.

  The spark plug 25 ignites the air-fuel mixture at the center of the combustion chamber 17 in the vicinity of the compression top dead center (reference numeral 6042). The spark plug 25 performs ignition after the compression top dead center, for example.

  As described above, by stratifying the air-fuel mixture, combustion noise can be suppressed and SPCCI combustion can be stabilized in the high load mid-rotation region (2).

(High load, low rotation range (3))
When the engine 1 is operating in the high load low rotation region (3), the engine 1 performs SI combustion instead of SPCCI combustion. The high load low rotation region (3) corresponds to the first high load region (C1) in the operation region map 701.

  Reference numeral 605 in FIG. 18 indicates fuel injection timing (reference numerals 6051 and 6052) and ignition timing (reference numeral 6053) when the engine 1 is operating in the operating state indicated by reference numeral 605 in the high load and low rotation region (3). An example of each combustion waveform (reference numeral 6054) is shown.

  The EGR system 55 introduces EGR gas into the combustion chamber 17 when the operating state of the engine 1 is in the high load low rotation region (3). The engine 1 reduces the amount of EGR gas as the load increases. At full open load, the EGR gas may be zero.

  When the engine 1 is operating in the high-load low-rotation region (3), 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. is there. The A / F of the air-fuel mixture may be set within the purification window of the three-way catalyst. Therefore, the excess air ratio λ of the air-fuel mixture may be set to 1.0 ± 0.2. By setting the air-fuel ratio of the air-fuel mixture to the stoichiometric air-fuel ratio, the fuel consumption performance is improved in the high load low rotation region (3). When the engine 1 is operated in the high load low rotation region (3), the fuel concentration of the entire air-fuel mixture in the combustion chamber 17 is 1 or less in the excess air ratio λ and the high load medium rotation region (2). The excess air ratio λ may be greater than or equal to the excess air ratio λ, preferably greater than the excess air ratio λ in the high load mid-rotation region (2).

  In the operation region map 702, when the engine 1 operates in the high-load low-rotation region (3), the injector 6 operates at the respective timings during the intake stroke and the retard period from the end of the compression stroke to the beginning of the expansion stroke. Then, fuel is injected into the combustion chamber 17 (reference numerals 6051 and 6052). By injecting fuel in two steps, the amount of fuel injected within the retard period can be reduced. By injecting fuel during the intake stroke (reference numeral 6051), it is possible to secure a sufficient time for forming the air-fuel mixture. Further, by injecting fuel during the retard period (reference numeral 6052), the flow in the combustion chamber 17 can be increased immediately before ignition, which is advantageous for stabilizing SI combustion.

  The spark plug 25 ignites the air-fuel mixture at a timing near the compression top dead center after fuel injection (reference numeral 6053). The spark plug 25 may ignite after compression top dead center, for example. The air-fuel mixture undergoes SI combustion in the expansion stroke. Since SI combustion starts in the expansion stroke, CI combustion does not start.

  When the engine 1 is operated in the high load low rotation region (3), the swirl flow is weaker than that in the high load medium rotation region (2). When operating in the high load low rotation region (3), the opening of the swirl control valve (SCV) 56 is larger than when operating in the high load medium rotation region (2). The opening degree of the swirl control valve 56 may be about 50% (that is, half open), for example.

  The axis of the injection hole of the injector 6 is displaced in the circumferential direction with respect to the spark plug 25. The fuel injected from the nozzle hole flows in the circumferential direction by the swirl flow in the combustion chamber 17. By the swirl flow, the fuel can be quickly transported to the vicinity of the spark plug 25. The fuel can be vaporized while being transported in the vicinity of the spark plug 25.

  On the other hand, if the swirl flow is too strong, the fuel is caused to flow in the circumferential direction, away from the vicinity of the spark plug 25, and the fuel cannot be quickly transported to the vicinity of the spark plug 25. Therefore, when the engine 1 is operated in the high load low rotation region (3), the swirl flow is weaker than that in the high load medium rotation region (2). As a result, the fuel can be quickly transported to the vicinity of the spark plug 25, so that the ignitability of the air-fuel mixture can be improved and the SI combustion can be stabilized.

(High rotation area (4))
When the rotational speed of the engine 1 is high, the time required for the crank angle to change by 1 ° is shortened. Therefore, for example, in the high rotation region in the high load region, it becomes difficult to stratify the air-fuel mixture in the combustion chamber 17 as described above. When the rotational speed of the engine 1 becomes high, it becomes difficult to perform the aforementioned SPCCI combustion.

  Therefore, when the engine 1 is operating in the high speed region (4), the engine 1 performs SI combustion instead of SPCCI combustion. The high rotation region (4) extends over the entire load direction from a low load to a high load.

  Reference numeral 606 in FIG. 18 indicates a fuel injection timing (reference numeral 6061) and an ignition timing (reference numeral 6062) and a combustion waveform when the engine 1 is operating in the operation state indicated by reference numeral 606 in the high speed region (4). (Reference numeral 6063) An example of each is shown.

  The EGR system 55 introduces EGR gas into the combustion chamber 17 when the operating state of the engine 1 is in the high rotation region (4). The engine 1 reduces the amount of EGR gas as the load increases. At full open load, the EGR gas may be zero.

  When the engine 1 is operated in the high speed region (4), the swirl control valve (SCV) 56 is fully opened. 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 in the high rotation region (4), and the pump loss can be reduced.

  When the engine 1 operates in the high speed region (4), 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. It is. The excess air ratio λ of the air-fuel mixture may be 1.0 ± 0.2. Note that the excess air ratio λ of the air-fuel mixture may be less than 1 in the high load region including the fully open load in the high rotation region (4).

  When the engine 1 operates in the high speed region (4), the injector 6 starts fuel injection during the intake stroke (see reference numeral 6061). The injector 6 injects fuel in a lump. By starting fuel injection during the intake stroke, a homogeneous or substantially homogeneous mixture can be formed in the combustion chamber 17. Further, when the engine 1 has a high rotation speed, the fuel vaporization time can be ensured as long as possible, so that unburned loss can be reduced and soot generation can be suppressed.

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

  Also in the engine 1 that employs the operation region map 702, it is preferable to adjust the SI rate as in the engine 1 that employs the operation region map 700. That is, as shown in FIG. 11, a region in which the SPCCI combustion is performed and the low load region (1) -1, the medium load region (1) -2, and the high load medium rotation region (2) are not supercharged. , The SI rate is increased linearly as the rotational speed of the engine 1 increases. By doing so, SPCCI combustion can be stably performed in a state where NVH, which may be a problem, is maintained below an allowable value even in a high rotation region.

  Also in the engine 1 that employs the operation region map 702, similarly to the engine 1 that employs the operation region map 700, the SI rate can be increased at the time of high rotation by the first to fourth means.

(Combustion waveform)
19 and 20 illustrate combustion waveforms in each operation state of the engine 1 adopting the operation region map 702. FIG. 19 is a simplified diagram corresponding to the operation region map 702 shown in FIG. 7C. The combustion waveform shows the change in the heat generation rate with respect to the change in the crank angle. W1 to W12 correspond to the respective operation states shown in FIG.

  As described above, in the region where SPCCI combustion is performed, the SI rate is increased as the rotational speed of the engine 1 increases. Comparing the waveforms of W2, W7, and W10, the peak of SI combustion gradually increases in the order of W2, W7, and W10. As a result, the peak of CI combustion gradually decreases in the order of W2, W7, and W10. Thereby, generation | occurrence | production of a combustion noise can be suppressed as the rotation speed of the engine 1 becomes high. In the region where SPCCI combustion is performed, the same tendency appears in W3, W8, W11 and W4, W9, W12, which have lower loads than W2, W7, and W10.

(Other embodiments)
The control of the engine 1 performed by the ECU 10 is not limited to the control based on the combustion model described above. For example, in the internal EGR system of the engine 1 described above, the internal EGR gas rate is changed by changing the negative overlap period, but may be controlled by opening the exhaust valve even during the intake stroke.

  The third and fourth means are not limited to supplementing the first and second means. Depending on the specification, it may be used independently of the first and second means. That is, the combination of the first to fourth means is arbitrary.

  Further, the engine 1 may perform SPCCI combustion over the entire rotation speed region, instead of performing SI combustion in the high rotation region. Further, the engine 1 may perform SPCCI combustion in the entire operation region.

  Further, the technology disclosed herein is not limited to being applied to the engine 1 having the above-described configuration. As the configuration of the engine 1, various configurations can be adopted.

1 Engine 10 ECU (controller)
17 Combustion chamber 23 Intake electric S-VT (internal EGR system)
24 Exhaust electric S-VT (internal EGR system)
25 Spark plug 3 Piston 31 Cavity 54 EGR valve (External EGR system)
55 EGR system 56 Swirl control valve (intake flow control device)
6 injector 102a SI rate changing means (heat quantity ratio changing means)
SW6 Acupressure sensor

Claims (8)

An engine configured to burn the air-fuel mixture in the combustion chamber;
A spark plug disposed facing the combustion chamber and configured to ignite an air-fuel mixture in the combustion chamber;
The burned gas is introduced into the combustion chamber by way of an intake passage that is attached to the engine and exhausts the burned gas generated in the combustion chamber from an exhaust passage through which the gas flows into the combustion chamber. An EGR system configured to:
A controller connected to each of the spark plug and the EGR system and configured to operate the engine by outputting a control signal to each of the spark plug and the EGR system;
With
The controller outputs a control signal to the spark plug at a predetermined ignition timing so that a predetermined form of combustion is performed in which the unburned mixture in the combustion chamber is combusted by self-ignition by ignition of the spark plug. ,
The controller controls the EGR system so that an EGR rate, which is a ratio of the amount of the burned gas contained in the air-fuel mixture in the combustion chamber, becomes lower when the engine speed is higher than when the engine speed is low. A control device for a compression ignition engine that outputs a signal. The control apparatus for a compression ignition engine according to claim 1,
An intake air flow control device attached to the engine and configured to change the flow of gas introduced into the combustion chamber;
The controller outputs a control signal to the EGR system so that the EGR rate is lower when the engine speed is higher than when the engine speed is low, and the intake flow control device is configured to increase the gas flow. A control device for a compression ignition engine that outputs a control signal. In the control device for a compression ignition engine according to claim 1 or 2,
The controller outputs a control signal to the EGR system so that the EGR rate becomes lower when the engine speed is higher than when the engine speed is low, and sends a control signal to the spark plug so that the ignition timing is advanced. Control device for compression ignition engine to output. In the control apparatus for the compression ignition engine according to any one of claims 1 to 3,
The controller is a control device for a compression ignition engine that outputs a control signal to the EGR system so that the EGR rate starts decreasing when the engine speed exceeds a preset reduction start speed . In the control apparatus for the compression ignition engine according to any one of claims 1 to 4,
A supercharging system attached to the engine and configured to supercharge gas introduced into the combustion chamber;
An internal EGR system provided in the engine and configured to change an internal EGR rate that is a ratio of an amount of internal EGR gas contained in an air-fuel mixture in the combustion chamber;
Further comprising
The controller performs supercharging when the supercharging system is in a first region where the load is higher than a predetermined load, and does not supercharge when the supercharging system is in a second region below the predetermined load. Output a control signal to the supercharging system,
The controller outputs a control signal to the EGR system so that the EGR rate is low when the engine speed is higher at least in the first region than when the engine speed is low, and in the second region, A control device for a compression ignition engine that outputs a control signal to the internal EGR system so that the internal EGR rate becomes high when the engine speed is higher than when the engine speed is low. In the control device for the compression ignition engine according to any one of claims 1 to 5,
The spark plug is disposed so as to face an upper central portion of the combustion chamber,
An internal EGR system provided in the engine and configured to change an internal EGR rate that is a ratio of the amount of internal EGR gas contained in the air-fuel mixture in the combustion chamber;
When the predetermined form of combustion is performed, a swirl flow is formed in the combustion chamber,
A control device for a compression ignition engine that outputs a control signal to the internal EGR system so that the internal EGR rate becomes high when the controller is higher than when the engine speed is low. In the control device for a compression ignition engine according to claim 5 or 6,
A control device for a compression ignition engine, in which the internal EGR gas is introduced into the combustion chamber when an intake valve opening period and an exhaust valve opening period overlap each other. The control apparatus for a compression ignition engine according to claim 7,
The controller outputs a control signal to the EGR system so as to start a decrease in the EGR rate when the engine speed exceeds a preset decrease start speed, and the load of the engine is A control device for a compression ignition engine that keeps the rate of decrease of the EGR rate substantially constant with respect to a change.

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