JP6528816B2 - Turbocharged compression self-ignition engine - Google Patents

Description

  The technology disclosed herein relates to a supercharged compression self-ignition engine.

  Patent Document 1 describes an engine that burns an air-fuel mixture in a combustion chamber by compression ignition in a predetermined region of low load and low rotation. This engine has a negative overlap period in which both the intake valve and the exhaust valve are closed by setting the closing timing of the exhaust valve before top dead center and the opening timing of the intake valve after top dead center in the predetermined region. Is provided. As a result, the high temperature EGR gas is introduced into the combustion chamber, the temperature in the combustion chamber becomes high, and the air-fuel mixture is compression-ignited.

  Patent Document 2 describes an engine in which a mechanical supercharger and an intercooler are provided in an intake passage, and the supercharged intake air is cooled and supplied to the engine. This engine deenergizes the mechanical supercharger by turning off the electromagnetic clutch in a predetermined operating condition, and closes the control valve of the air bypass passage bypassing the mechanical supercharger and the inter , Pre-rotation mechanical supercharger by pressure difference. This prevents the occurrence of torque shock when the electromagnetic clutch is turned on.

JP, 2009-197740, A Patent No. 3564989

  In combustion by compression ignition, the temperature in the combustion chamber must be adjusted to an appropriate range. For example, when the outside air temperature is high, the compression end temperature which is the temperature in the combustion chamber when the compression top dead center is reached becomes too high, and the combustion by compression ignition becomes steep and the combustion noise exceeds the allowable value.

  The technology disclosed herein adjusts the combustion chamber to an appropriate temperature in a supercharged compression self-ignition engine.

  Specifically, the compression self-ignition type engine with a supercharger disclosed herein includes an intake passage connected to a combustion chamber of the engine, a supercharger disposed in the intake passage, and a drive of the supercharger. A switching unit for switching the motor and the non-drive, an intercooler disposed downstream of the supercharger in the intake passage, and a bypass passage connecting the upstream of the supercharger in the intake passage and the downstream of the intercooler A flow control valve disposed in the bypass passage, a variable valve mechanism configured to change the opening / closing operation of an intake valve interposed between the downstream end of the intake passage and the combustion chamber; A control unit connected to the switching unit, the flow control valve, and the variable valve mechanism, and outputting a control signal to the switching unit, the flow control valve, and the variable valve mechanism; Said fuel when Comprising a temperature estimation unit that estimates a compression end temperature which is the temperature of the chamber, the.

  When the control unit operates the engine without driving the supercharger through the switching unit, the control unit opens the flow control valve, and in the engine, the air-fuel mixture in the combustion chamber is burned by compression self-ignition Do.

  The control unit also causes the compression end temperature to be a preset allowable temperature based on the compression end temperature estimated by the temperature estimation unit when the engine is operated with the supercharger not driven. And when the engine compression ratio is lower than that when the compression end temperature is below the allowable temperature, the closing timing of the intake valve is changed through the variable valve mechanism.

  Here, the “engine” may be a four-stroke engine which is operated by the combustion chamber repeating an intake stroke, a compression stroke, an expansion stroke and an exhaust stroke.

  When the engine is operated without driving the supercharger, the flow control valve in the bypass passage is opened. The supercharger may be a mechanical supercharger driven by an engine. Moreover, it is good also as an electrically driven supercharger driven by an electrical energy. When the turbocharger is not driven, the intake air passes through the bypass passage and enters the combustion chamber without passing through the turbocharger and the intercooler. On the other hand, when the engine is operated while the supercharger is in operation, intake air is boosted by the supercharger and cooled by the intercooler to enter the combustion chamber. When the supercharger is driven, the supercharging pressure of the engine may be adjusted by appropriately adjusting the opening degree of the flow control valve.

  When the engine operates in a state where the supercharger is not driven, the mixture in the combustion chamber burns due to compression self-ignition. The combustion chamber must be adjusted to an appropriate temperature. When the compression end temperature exceeds a preset allowable temperature, the closing timing of the intake valve is changed through the variable valve mechanism so that the effective compression ratio of the engine is lower than when the compression end temperature is lower than the allowable temperature. For example, when the closing timing of the intake valve is delayed after the bottom dead center, the effective compression ratio of the engine decreases. The change of the closing timing of the intake valve is not limited to changing the phase of the valve timing of the intake valve. For example, by changing the lift amount of the intake valve or changing the valve opening angle of the intake valve (that is, the length of the period during which the intake valve is open), the closing timing of the intake valve is changed. It is also good.

  By lowering the effective compression ratio of the engine, the temperature rise due to the gas in the combustion chamber being compressed during the compression stroke decreases, so the temperature at the compression end decreases accordingly. It is possible to keep the compression end temperature below the allowable temperature. The temperature estimation unit may estimate the compression end temperature based on detection values of various sensors.

  Here, it is conceivable to lower the temperature of the intake air supplied to the combustion chamber using, for example, an intercooler when the compression end temperature rises. However, if the supercharger is driven so that the intake air passes through the intercooler, the fuel efficiency performance of the engine is reduced.

  On the other hand, in the above-described configuration in which the closing timing of the intake valve is changed to lower the effective compression ratio of the engine, the compression end temperature can be made equal to or lower than the allowable temperature without lowering the fuel efficiency performance of the engine. As a result, when the engine is operated in a state where the supercharger is not driven, the air-fuel mixture in the combustion chamber can be appropriately burned by compression self-ignition.

  The control unit deactivates the supercharger when the load on the engine is low, and the variable valve mechanism includes a mechanism for changing the phase of the valve timing of the intake valve, and the load on the engine When the engine speed is low, the EGR gas introduced into the combustion chamber is set to a predetermined ratio or more by providing an overlap period in which both the intake valve and the exhaust valve are open, and the control unit also performs the supercharging When the compression end temperature exceeds the allowable temperature while the engine is operating with the machine not driven, the overlap period becomes shorter than when the compression end temperature is lower than the allowable temperature and The phase of the valve timing of the intake valve may be retarded through the variable valve mechanism so that the effective compression ratio of the engine decreases.

  The temperature in the combustion chamber can be increased by providing an overlap period and introducing an EGR gas at a predetermined ratio or more into the combustion chamber. When the load on the engine is low, combustion by compression self-ignition can be stabilized. When the compression end temperature exceeds the allowable temperature, retarding the phase of the valve timing of the intake valve results in a shorter overlap period than when the compression end temperature is lower than the allowable temperature. As a result, the amount of EGR gas introduced into the combustion chamber decreases, so the temperature in the combustion chamber decreases before starting the compression. In addition, when the closing timing of the intake valve is delayed after the bottom dead center by retarding the phase of the valve timing of the intake valve, the effective compression ratio of the engine decreases.

  Therefore, the above configuration has two effects of lowering the temperature in the combustion chamber before starting compression and retarding the effective compression ratio of the engine by retarding the phase of the valve timing of the intake valve. can get. As a result, the compression end temperature can be effectively reduced and the compression end temperature can be suppressed to the allowable temperature or less, so that the mixture in the combustion chamber can be burned by compression self-ignition appropriately.

When the compression end temperature exceeds the allowable temperature and the outside air temperature exceeds a predetermined temperature, the control unit deactivates the supercharger through the switching unit, and the compression end temperature is equal to the compression end temperature. The flow control valve is closed more than when the temperature is lower than the allowable temperature, and the control unit is configured to compress the engine effectively when the outside temperature is lower than the predetermined temperature when the compression end temperature exceeds the allowable temperature. the ratio of the so compression end temperature is lower than in the following the permissive temperature to change the closing timing of the intake valve via the variable valve mechanism.

  When the outside air temperature exceeds the predetermined temperature and is too high, the compression end temperature may not be able to be suppressed to the allowable temperature or less depending on the above-described configuration for reducing the effective compression ratio of the engine. Therefore, when the outside air temperature exceeds the predetermined temperature, the temperature of the intake air is reduced using the intercooler. Specifically, the turbocharger is not driven and the flow control valve of the bypass passage is closed. As a result, a pressure difference is generated before and after the turbocharger, and the turbocharger rotates (that is, pre-rotations) due to the pressure difference, and the intake air passes through the turbocharger. The intake air passing through the supercharger enters the combustion chamber after being cooled by the intercooler. By adjusting the opening degree of the flow control valve in the bypass passage, the pressure difference between the front and back of the turbocharger changes, and the ratio of the intake amount passing through the turbocharger and the intercooler to the intake amount passing through the bypass passage becomes change. Since the temperature of the intake air introduced into the combustion chamber can be adjusted, the compression end temperature can be suppressed to the allowable temperature or less even when the outside air temperature is too high. Although this arrangement does not drive the supercharger, pump losses are somewhat increased.

  On the other hand, when the outside air temperature is equal to or lower than the predetermined temperature, the compression end temperature is suppressed to the allowable temperature or less by lowering the effective compression ratio of the engine as described above. This configuration is advantageous in terms of the fuel consumption performance of the engine because it does not drive the turbocharger and does not increase the pump loss.

  The control unit drives the supercharger through the switching unit when the compression end temperature exceeds the allowable temperature and the outside air temperature exceeds a second predetermined temperature higher than the predetermined temperature. The flow control valve may be opened.

  This configuration is a configuration that enables effective cooling of the intake air using the intercooler when the outside air temperature is further high. That is, when driving the supercharger and opening the flow control valve, part of the intake air whose temperature is lowered by passing through the supercharger and the intercooler enters the combustion chamber, and the rest is through the bypass passage. It is returned to the upstream of the turbocharger. The intake air returned upstream of the supercharger passes the supercharger and the intercooler again with fresh intake. Part of the intake air whose temperature has been reduced by passing the turbocharger and the intercooler enters the combustion chamber, and the remainder is returned to the upstream of the turbocharger again through the bypass passage. Thus, a part of the intake air can be circulated in the circulation passage formed by the intake passage provided with the intercooler and the bypass passage, so that the intercooler sufficiently cools even when the outside air temperature is high. Can be introduced into the combustion chamber. As a result, the compression end temperature can be suppressed to the allowable temperature or less.

  The engine may have a geometric compression ratio of 13 or more and 30 or less.

  If the geometric compression ratio of the engine is relatively high, which is advantageous for stabilization of combustion by compression self-ignition, it may lead to pre-ignition. In the engine having a geometric compression ratio of 13 or more and 30 or less, by adjusting the compression end temperature as described above, it is possible to stabilize the combustion by the compression self-ignition while avoiding the pre-ignition.

  In addition to the advantage of reducing the cooling loss by not increasing the geometric compression ratio too much, it is also advantageous to reduce the mechanical loss. Fuel efficiency of the engine is improved.

  The fuel supplied to the combustion chamber may include at least gasoline.

  Fuel containing gasoline may cause pre-ignition in a high temperature combustion chamber, but as described above, by adjusting the compression end temperature, the engine using the fuel containing gasoline is prematurely removed. The combustion by the compression self-ignition can be stabilized while avoiding the ignition.

  As described above, according to the above-described compression self-ignition engine with a supercharger, the temperature in the combustion chamber can be adjusted to an appropriate temperature.

FIG. 1 is a diagram illustrating the configuration of an engine. FIG. 2 is a view illustrating the configuration of the combustion chamber, and the upper view is a plan view equivalent view of the combustion chamber, and the lower portion is a cross-sectional view taken along the line AA. FIG. 3 is a plan view illustrating the configuration of the combustion chamber and the intake system. The left diagram of FIG. 4 illustrates the configuration of the main circuit of the engine cooling system, and the right diagram of FIG. 4 illustrates the configuration of the engine cooling system subcircuit. FIG. 5 is a block diagram illustrating the configuration of a control device of an engine. FIG. 6 is a diagram illustrating an operating range map of the engine. FIG. 7 is a diagram illustrating the fuel injection timing and the ignition timing in each operation region and the combustion waveform. FIG. 8 is a diagram illustrating a rig test apparatus for measuring a swirl ratio. FIG. 9 is a diagram illustrating the relationship between the opening ratio of the secondary passage and the swirl ratio. FIG. 10 is a diagram showing the flow of gas in the intake passage when driving the turbocharger and performing circulation control of the turbocharger. FIG. 11 is a view corresponding to FIG. 10 showing the flow of gas in the intake passage when the turbocharger is not driven. FIG. 12 is a view corresponding to FIG. 10 showing the flow of gas in the intake passage when performing prerotation control of the turbocharger. FIG. 13 is a diagram illustrating the relationship between the flow rate ratio of intake air passing through the intercooler and the gas temperature in the surge tank when performing the prerotation control of the supercharger. FIG. 14 is a diagram illustrating changes in in-cylinder pressure and changes in in-cylinder temperature when the closing timing of the intake valve is retarded. FIG. 15 is a flowchart illustrating the control process of the engine according to the adjustment of the compression end temperature. FIG. 16 is a part of a flowchart illustrating a reference example of a control process of an engine.

  Hereinafter, an embodiment of a supercharged compression self-ignition engine will be described in detail based on the drawings. The following description is an example of an engine. FIG. 1 is a diagram illustrating the configuration of an engine. FIG. 2 is a diagram illustrating the configuration of the combustion chamber. FIG. 3 is a diagram illustrating the configuration of the combustion chamber and the intake system. The intake side in FIG. 1 is on the left side in the drawing, and the exhaust side is on the right side in the drawing. The intake side in FIGS. 2 and 3 is on the right side in the drawing, and the exhaust side is on the left in the drawing. FIG. 4 is a diagram illustrating the configuration of a cooling device of an engine. FIG. 5 is a block diagram illustrating the configuration of a control device of an engine.

  The engine 1 is a four-stroke engine operated by the combustion chamber 17 repeating an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke. The engine 1 is mounted on a four-wheeled vehicle. By driving the engine 1, the car travels. The fuel of the engine 1 is gasoline in this configuration example. The fuel may be gasoline including bioethanol and the like. The fuel of the engine 1 may be any fuel as long as it is a liquid fuel containing at least gasoline.

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

  A piston 3 is slidably inserted in each cylinder 11. The piston 3 is connected to the crankshaft 15 via a connecting rod 14. The piston 3 defines the 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 the 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.

  The lower surface of the cylinder head 13, that is, the ceiling surface of the combustion chamber 17, is constituted by an inclined surface 1311 and an inclined surface 1312 as shown in the upper view of FIG. The inclined surface 1311 has an upward slope toward the injection axis X2 of the injector 6 described later from the intake side. The inclined surface 1312 has an upward slope from the exhaust side toward the injection axis X2. The ceiling surface of the combustion chamber 17 has a so-called pent roof shape.

  The upper surface of the piston 3 is raised toward the ceiling surface of the combustion chamber 17. A cavity 31 is formed on the upper surface of the piston 3. The cavity 31 is recessed from the upper surface of the piston 3. The cavity 31 faces the injector 6 described later.

  The center of the cavity 31 is shifted to the exhaust side relative to the central axis X1 of the cylinder 11. The center of the cavity 31 coincides with the injection axis X 2 of the injector 6. The cavity 31 has a convex portion 311. The convex portion 311 is provided on the injection axis X2 of the injector 6. The convex portion 311 is substantially conical. The convex portion 311 extends upward from the bottom of the cavity 31 toward the ceiling surface of the combustion chamber 17.

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

  The circumferential side surface of the recessed portion 312 is inclined from the bottom surface of the cavity 31 toward the opening of the cavity 31 with respect to the injection axis X2. The inner diameter of the cavity 31 in the recess 312 gradually increases from the bottom of the cavity 31 toward the opening of the cavity 31.

  The geometric compression ratio of the engine 1 is set to 13 or more and 30 or less. As described later, the engine 1 performs SPCCI combustion in which SI combustion and CI combustion are combined in a part of the operation range. SPCCI combustion controls CI combustion using heat generation and pressure increase due to SI combustion. In the engine 1, it is not necessary to increase the temperature of the combustion chamber 17 (that is, the compression end temperature) when the piston 3 reaches compression top dead center for self-ignition of air-fuel mixture. That is, although the engine 1 performs CI combustion, it is possible to set its geometric compression ratio relatively low. Lowering the geometric compression ratio is advantageous for reducing cooling loss and reducing mechanical loss. The geometric compression ratio of the engine 1 may be 14 to 17 in the regular specification (the fuel octane number is about 91), and may be 15 to 18 in the high fuel specification (the fuel octane number is about 96).

  An intake port 18 is formed in the cylinder head 13 for each cylinder 11. The intake port 18 has two intake ports, a first intake port 181 and a second intake port 182, as shown in FIG. The first intake port 181 and the second intake port 182 are aligned in the axial direction of the crankshaft 15, that is, in the front-rear direction of the engine 1. The intake port 18 is in communication with the combustion chamber 17. The intake port 18 is a so-called tumble port, which is not shown in detail. That is, the intake port 18 is shaped such that a tumble flow is formed in the combustion chamber 17.

  An intake valve 21 is disposed at the intake port 18. The intake valve 21 opens and closes between the combustion chamber 17 and the intake port 18. The intake valve 21 is opened and closed at a predetermined timing by a valve operating mechanism. The valve operating mechanism may be a variable valve operating mechanism that makes the valve timing and / or the valve lift variable. In this configuration example, as shown in FIG. 5, the variable valve mechanism has an intake electric motor S-VT (Sequential-Valve Timing) 23. The intake electric motor S-VT 23 is configured to continuously change the rotational phase of the intake camshaft within a predetermined angular range. Thereby, the opening timing and closing timing of the intake valve 21 continuously change. The intake valve mechanism may have a hydraulic S-VT instead of the electric S-VT.

  Further, an exhaust port 19 is formed in the cylinder head 13 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 aligned 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 at the exhaust port 19. The exhaust valve 22 opens and closes between the combustion chamber 17 and the exhaust port 19. The exhaust valve 22 is opened and closed at a predetermined timing by a valve operating mechanism. The valve operating mechanism may be a variable valve operating mechanism that varies valve timing and / or valve lift. In this configuration example, as shown in FIG. 5, the variable valve mechanism has an exhaust motor drive S-VT 24. The exhaust motor S-VT 24 is configured to continuously change the rotational phase of the exhaust camshaft within a predetermined angular range. Thereby, the opening timing and closing timing of the exhaust valve 22 change continuously. The exhaust valve mechanism may have a hydraulic S-VT instead of the electric S-VT.

  The 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 motor S-VT 23 and the exhaust electric motor S-VT 24. By this, the residual gas in the combustion chamber 17 is scavenged. In addition, the internal EGR (Exhaust Gas Recirculation) gas is introduced into the combustion chamber 17 or confined in the combustion chamber 17 by adjusting the length of the overlap period. In this configuration example, the intake motor S-VT 23 and the exhaust motor S-VT 24 constitute an internal EGR system. The internal EGR system is not necessarily configured by the S-VT.

  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 the valley portion of the pent roof where the inclined surface 1311 on the intake side and the inclined surface 1312 on the exhaust side intersect. As shown in FIG. 2, the injector axis X 2 of the injector 6 is disposed on the exhaust side of the central axis X 1 of the cylinder 11. The injection axis X2 of the injector 6 is parallel to the central axis X1. The injection axis X2 of the injector 6 coincides with the position of the convex portion 311 of the cavity 31 as described above. The injector 6 is opposed to the cavity 31. The injection axis X2 of the injector 6 may be coincident with the central axis X1 of the cylinder 11. Also in this case, it is desirable that the injection axis X2 of the injector 6 and the position of the convex portion 311 of the cavity 31 be coincident with each other.

  The injector 6 is constituted by a multi-injection-type fuel injection valve having a plurality of injection holes although detailed illustration is omitted. The injector 6 injects fuel so that the fuel spray spreads radially from the center of the combustion chamber 17 as indicated by a two-dot chain line in FIG. In the present configuration example, the injector 6 has ten injection holes, and the injection holes are disposed equiangularly in the circumferential direction. The axis of the injection hole is circumferentially offset with respect to the spark plug 25 described later, as shown in the upper view of FIG. That is, the spark plug 25 is sandwiched between the axes of two adjacent injection holes. This prevents the fuel spray injected from the injector 6 from directly hitting the spark plug 25 and wetting the electrode.

  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 passage 62 connecting the fuel tank 63 and the injector 6 to each other. A fuel pump 65 and a common rail 64 are interposed in the fuel supply passage 62. The fuel pump 65 pumps fuel to the common rail 64. The fuel pump 65 is a plunger type pump driven by the crankshaft 15 in this configuration example. The common rail 64 is configured to store the fuel pumped by 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 from the injection port of the injector 6 into the combustion chamber 17. The fuel supply system 61 is configured to be able to supply fuel with a high pressure of 30 MPa or more to the injector 6. The maximum fuel pressure of the fuel supply system 61 may be, for example, about 120 MPa. The pressure of the fuel supplied to the injector 6 may be changed according to the operating state of the engine 1. The configuration of the fuel supply system 61 is not limited to the above configuration.

  A spark plug 25 is attached to the cylinder head 13 for each cylinder 11. The spark plug 25 forcibly ignites the mixture in the combustion chamber 17. The spark plug 25 is disposed closer to the intake side than the central axis X1 of the cylinder 11 in this configuration example. The spark plug 25 is located between the two intake ports 18. The spark plug 25 is attached to the cylinder head 13 in such a way as to approach the center of the combustion chamber 17 from the upper side to the lower side. The electrode of the spark plug 25 faces the combustion chamber 17 and is located near the ceiling surface of the combustion chamber 17 as shown in FIG. The spark plug 25 may be disposed closer to the exhaust side than the central axis X1 of the cylinder 11. Further, while the spark plug 25 is disposed on the central axis X1 of the cylinder 11, the injector 6 may be disposed on the intake side or the exhaust side of the central axis X1 of the cylinder 11.

  An intake passage 40 is connected to one side of the engine 1. The intake passage 40 communicates with the intake port 18 of each cylinder 11. The intake passage 40 is a passage through which the gas introduced into the combustion chamber 17 flows. An air cleaner 41 for filtering fresh air is disposed at the upstream end of the intake passage 40. In the vicinity of the downstream end of the intake passage 40, a surge tank 42 is disposed. 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 adjust the introduction amount of fresh air into the combustion chamber 17 by adjusting the opening degree of the valve.

  A supercharger 44 is disposed downstream of the throttle valve 43 in the intake passage 40. The supercharger 44 is configured to supercharge the gas introduced into the combustion chamber 17. In this configuration example, the turbocharger 44 is a mechanical turbocharger driven by the engine 1. The mechanical supercharger 44 may be, for example, of the Richolem type. The configuration of the mechanical turbocharger 44 may be any configuration. The mechanical supercharger 44 may be of the roots type, vane type, or centrifugal type.

  An electromagnetic clutch 45 is interposed between the turbocharger 44 and the engine 1. The electromagnetic clutch 45 transmits driving power from the engine 1 to the turbocharger 44 and blocks transmission of driving power between the turbocharger 44 and the engine 1. As will be described later, the supercharger 44 is switched on and off as the ECU 10 switches between disconnection and connection of the electromagnetic clutch 45. The engine 1 can switch between supercharging of the gas introduced into the combustion chamber 17 by the supercharger 44 and noncharging of the gas introduced into the combustion chamber 17 by the supercharger 44. It is configured.

  An intercooler 46 is disposed downstream of the turbocharger 44 in the intake passage 40. The intercooler 46 is configured to cool the gas compressed at the turbocharger 44. The intercooler 46 may be, for example, water-cooled. The intercooler 46 may be oil-cooled.

  A bypass passage 47 is connected to the intake passage 40. The bypass passage 47 connects the upstream portion of the turbocharger 44 and the downstream portion of the intercooler 46 with each other in the intake passage 40 so as to bypass the turbocharger 44 and the intercooler 46. More specifically, the bypass passage 47 is connected to the surge tank 42. An air bypass valve 48 is disposed in the bypass passage 47. The air bypass valve 48 regulates the flow rate of gas flowing through the bypass passage 47.

  When the turbocharger 44 is turned off (that is, when the electromagnetic clutch 45 is disconnected), the air bypass valve 48 is fully opened. Thus, the gas flowing through the intake passage 40 bypasses the turbocharger 44 and is introduced into the combustion chamber 17 of the engine 1. The engine 1 operates in a non-supercharged state, ie, in a state of natural intake.

  When the turbocharger 44 is turned on (that is, when the electromagnetic clutch 45 is connected), part of the gas that has passed through the turbocharger 44 flows back through the bypass passage 47 upstream of the turbocharger 44. Do. Since the reverse flow rate can be adjusted by adjusting the opening degree of the air bypass valve 48, the supercharging pressure of the gas introduced into the combustion chamber 17 can be adjusted. The supercharging time is defined as the time when the pressure in the surge tank 42 exceeds the atmospheric pressure, and the non-supercharging time is defined as the time when the pressure in the surge tank 42 becomes lower than the atmospheric pressure. It is also good.

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

  The engine 1 has a swirl generating portion that generates a swirl flow in the combustion chamber 17. The swirl generating portion is a swirl control valve 56 attached to the intake passage 40, as shown in FIG. 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 capable of reducing the cross section of the secondary passage. When the degree of opening of the swirl control valve 56 is small, of 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 flow rate of intake air flowing from the second intake port 182 into the combustion chamber 17 is relatively reduced, so the swirl flow in the combustion chamber 17 becomes strong. When the degree of 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 even, so the swirl flow in the combustion chamber 17 is weak. Become. When the swirl control valve 56 is fully opened, no swirl flow occurs. The swirl flow circulates in the counterclockwise direction in FIG. 3 as shown by the white arrow (see also the white arrow in FIG. 2).

  The swirl generating part shifts the open period of the two intake valves 21 instead of attaching the swirl control valve 56 to the intake passage 40 or in addition to attaching the swirl control valve 56, and one intake valve 21 A configuration in which intake air can be introduced into the combustion chamber 17 from only may be adopted. By opening only one of the two intake valves 21, the intake air is unequally introduced into the combustion chamber 17, so that swirl flow can be generated in the combustion chamber 17. . Furthermore, the swirl generating portion may be configured to generate a swirl flow in the combustion chamber 17 by devising the shape of the intake port 18.

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

  The exhaust passage 50 is provided with an exhaust gas purification system having a plurality of catalytic converters. The upstream catalytic converter, which is not shown, 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 of the illustrated example. For example, GPF may be omitted. Also, 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 recirculating a part of the 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 of the turbocharger 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 turbocharger 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 adjust the flow rate of the burnt gas flowing through the EGR passage 52. By adjusting the opening degree of the EGR valve 54, it is possible to adjust the reflux amount of the cooled burned gas, that is, the external EGR gas.

  In this configuration example, the EGR system 55 includes an external EGR system configured to include the EGR passage 52 and the EGR valve 54, and an interior configured to include the intake electric motor S-VT 23 and the exhaust motor S-VT 24 described above. And an EGR system. The EGR valve 54 also constitutes one of the state quantity setting devices. In the external EGR system, since the EGR passage 52 is connected downstream of the GPF 512 and the EGR cooler 53 is provided, burnt gas having a temperature lower than that of the internal EGR system can be supplied to the combustion chamber 17 it can.

  As shown in FIG. 4, the cooling device 71 of the engine 1 includes a main circuit 71A and a sub circuit 71B. The main circuit 71A and the sub circuit 71B are independent of each other. That is, the refrigerant does not move back and forth between the main circuit 71A and the sub circuit 71B.

  The main circuit 71A is a variable displacement type that supplies the engine cooled by the main radiator 72, which cools the refrigerant using traveling wind, and the refrigerant cooled by the main radiator 72, to the engine body 100 including the cylinder head 13 and the cylinder block 12. The water pump 74 of The water pump 74 is driven by the engine 1. The refrigerant supplied to the engine main body 100 cools each part of the engine main body 100 and then is discharged from the engine main body 100 and returns to the main radiator 72.

  Similar to the main radiator 72, the sub circuit 71B includes a sub radiator 75 for cooling the refrigerant using traveling air, and an electric water pump 76 for supplying the refrigerant cooled by the sub radiator 75 to the intercooler 46. ing. The refrigerant supplied to the intercooler 46 cools the gas passing through the intercooler 46, and then is discharged from the intercooler 46 and returns to the sub radiator 75.

  The refrigerant flowing through the main circuit 71A passes through the inside of the engine body 100. The refrigerant flowing through the sub circuit 71B does not pass through the inside of the engine body 100. Therefore, the refrigerant flowing through the main circuit 71A has a higher temperature than the refrigerant flowing through the sub circuit 71B. The EGR cooler 53 is connected to the main circuit 71A, although not shown.

  The control device of the compression self-ignition type engine includes an ECU (Engine Control Unit) 10 for operating the engine 1. The ECU 10 is a controller based on a well-known microcomputer, and as shown in FIG. 5, a central processing unit (CPU) 101 that executes a program and, for example, a random access memory (RAM) or a ROM. A memory 102 configured by (Read Only Memory) to store programs and data, and an input / output bus 103 for inputting / outputting an electric signal are provided. The ECU 10 is an example of a control unit.

  As shown in FIGS. 1 and 5, various sensors SW1 to SW16 are connected to the ECU 10. The sensors SW1 to SW16 output detection signals to the ECU 10. The sensors include the following sensors.

That is, an air flow sensor SW1 disposed downstream of the air cleaner 41 in the intake passage 40 and detecting the flow rate of fresh air flowing through the intake passage 40; a first intake temperature sensor SW2 detecting the temperature of the fresh air; The first pressure sensor SW3 disposed downstream of the connection position of the EGR passage 52 at 40 and upstream of the turbocharger 44 and detecting the pressure of the gas flowing into the turbocharger 44; A second intake air temperature sensor SW4 disposed downstream of the engine 44 and upstream of the connection position of the bypass passage 47 and detecting the temperature of the gas flowing out of the turbocharger 44, attached to the surge tank 42, and A second pressure sensor SW5 for detecting the pressure of gas downstream of the feeder 44, attached to the cylinder head 13 corresponding to each cylinder 11, and A finger pressure sensor SW6 for detecting the pressure in the burning chamber 17, an exhaust temperature sensor SW7 disposed in the exhaust passage 50 and detecting the temperature of the exhaust gas discharged from the combustion chamber 17, upstream of the catalytic converter upstream in the exhaust passage 50 arranged and the linear O ₂ sensor SW8 for detecting the oxygen concentration in the exhaust gas, and is disposed downstream of the three-way catalyst 511 upstream of the catalytic converter, the lambda O ₂ sensor SW9 for detecting the oxygen concentration in the exhaust gas A coolant temperature sensor SW10 attached to the engine 1 for detecting the temperature of the cooling water, a crank angle sensor SW11 attached to the engine 1 for detecting the rotation angle of the crankshaft 15, attached to an accelerator pedal mechanism and an accelerator pedal Accelerator opening sensor SW1 that detects the accelerator opening corresponding to the amount of operation 2. Intake cam angle sensor SW13 attached to the engine 1 and detecting the rotational angle of the intake camshaft, exhaust cam angle sensor SW14 attached to the engine 1 and detecting the rotational angle of the exhaust camshaft, to the EGR passage 52 An EGR differential pressure sensor SW15 disposed and detecting a differential pressure upstream and downstream of the EGR valve 54, and a fuel pressure sensor attached to the common rail 64 of the fuel supply system 61 and detecting the pressure of the fuel supplied to the injector 6 It is SW16.

  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 controls the control signal related to the calculated control amount as the injector 6, spark plug 25, intake motor S-VT 23, exhaust motor S-VT 24, fuel supply system 61, throttle valve 43, EGR valve 54, supercharger 44. The electromagnetic clutch 45, the air bypass valve 48, and the swirl control valve 56 are output.

  For example, the ECU 10 sets the target torque of the engine 1 and determines the target boost pressure based on the detection signal of the accelerator opening sensor SW12 and the map set in advance. Then, the ECU 10 adjusts the opening degree of the air bypass valve 48 based on the target boost pressure and the differential pressure across the turbocharger 44 obtained from the detection signals of the first pressure sensor SW3 and the second pressure sensor SW5. Thus, the feedback control is performed such that the boost pressure becomes the target boost pressure.

  Further, the ECU 10 sets a target EGR rate (that is, the ratio of the EGR gas to the total gas in the combustion chamber 17) based on the operating state of the engine 1 and a map set in advance. Then, the ECU 10 determines the target EGR gas amount based on the target EGR rate and the intake air amount based on the detection signal of the accelerator opening sensor SW12, and also before and after the EGR valve 54 obtained from the detection signal of the EGR differential pressure sensor SW15. By adjusting the opening degree of the EGR valve 54 based on the differential pressure, feedback control is performed so that the amount of external EGR gas introduced into the combustion chamber 17 becomes the target amount of EGR gas.

Furthermore, the ECU 10 performs air-fuel ratio feedback control when a predetermined control condition is satisfied. Specifically ECU10 includes a linear O ₂ sensor SW8 and, based on the oxygen concentration in the exhaust gas detected by the lambda O ₂ sensor SW9, as the air-fuel ratio of the mixture has a desired value, the fuel injector 6 Adjust the injection amount.

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

(Operating area of engine)
FIG. 6 illustrates an operating area map of the engine 1 at warm time. The operating area map 501, 502 of the engine 1 is determined by the load and the rotational speed, and is divided into five areas for high and low of the load and high and low of the rotational speed. Specifically, the five regions include idle operation, and the low load region (1) -1 extending to the low rotation and medium rotation regions, higher in load than the low load region, and the low rotation and medium rotation regions Medium load area (1) -2 which spreads to the middle load area (1) -2 and in which the load is higher than the middle load area including the full open load (2) middle load area in the high load area Low rotation area (3) whose rotation speed is lower than (2), and low load area (1) -1, medium load area (1) -2, high load medium rotation area (2), and high load low It is a high rotation area (4) whose rotation speed is higher than that of the rotation area (3). Here, when the low rotation region, the middle rotation region, and the high rotation region are all divided into three regions of the low rotation region, the middle rotation region, and the high rotation region in the rotational speed direction, respectively. The low rotation region, the middle rotation region, and the high rotation region may be used. In the example of FIG. 6, the number of rotations less than N1 is low, the number of rotations N2 or more is high, and the number of rotations N1 or more and N2 is medium. The rotational speed N1 may be, for example, about 1200 rpm, and the rotational speed N2 may be, for example, about 4000 rpm. The two-dot chain line in FIG. 6 indicates the load-load line (Road-Load Line) of the engine 1. In FIG. 6, the operation area maps 501 and 502 of the engine 1 are drawn in two parts for ease of understanding. The map 501 shows the state of air-fuel mixture and combustion mode in each area, and the driving area and the non-driving area of the turbocharger 44. The map 502 shows the opening degree of the swirl control valve 56 in each area.

  The engine 1 is compressed in a low load range (1) -1, a medium load range (1) -2, and a high load medium rotation range (2) mainly for the purpose of improving the fuel efficiency and the exhaust gas performance. It burns by ignition. The engine 1 also performs spark ignition combustion in other regions, specifically, a high load low rotation region (3) and a 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. 7, reference numerals 601, 602, 603, 604, 605 and 606 respectively correspond to the operating state of the engine 1 indicated by reference numerals 601, 602, 603, 604, 605 and 606 in the operating area map 501 of FIG. .

(Low load area (1) -1)
When the engine 1 is operating in the low load range (1) -1, the engine 1 performs CI combustion. In the self-ignition combustion, when the temperature in the combustion chamber 17 before the start of compression varies, the timing of the self-ignition changes significantly. Therefore, the engine 1 performs SPCCI combustion in which SI combustion and CI combustion are combined in the low load range (1) -1.

  Reference numeral 601 in FIG. 7 indicates the fuel injection timing (reference numerals 6011 and 6012) and the ignition timing (reference numeral 6013) when the engine 1 is operating in the operation state 601 in the low load range (1) -1. A waveform (ie, a waveform representing a change in heat release rate with respect to the crank angle, symbol 6014) is shown.

  In the SPCCI combustion, the ignition plug 25 forcibly ignites the mixture in the combustion chamber 17 so that the mixture causes SI combustion by flame propagation, and the heat of SI combustion generates heat in the combustion chamber 17. As the temperature rises and the pressure in the combustion chamber 17 rises due to flame propagation, the unburned air-fuel mixture performs CI combustion by self-ignition.

  By adjusting the amount of heat generation of SI combustion, it is possible to absorb temperature variations in the combustion chamber 17 before the start of compression. Even if the temperature in the combustion chamber 17 before the start of compression varies, for example, if the start timing of SI combustion is adjusted by adjusting the ignition timing, the air-fuel mixture can be self-ignited at the target timing.

  When performing the SPCCI combustion, the spark plug 25 ignites the air-fuel mixture near the compression top dead center, whereby combustion by flame propagation starts. Heat generation during SI combustion is milder than heat generation during CI combustion. Therefore, the heat generation rate waveform 6014 has a relatively small rising slope. In addition, pressure fluctuation (dp / dθ) in the combustion chamber 17 also becomes gentler during SI combustion than during CI combustion.

  As the temperature and pressure in the combustion chamber 17 increase due to SI combustion, the unburned mixture self-ignites. In the example of FIG. 7, the slope of the heat generation rate waveform 6014 changes from small to large at the timing of self-ignition. That is, the waveform of the heat release rate has an inflection point at the timing when the CI combustion starts.

  After the start of CI combustion, SI combustion and CI combustion are performed in parallel. The rate of heat release is relatively large because CI combustion generates heat more than SI combustion. However, since the CI combustion is performed after the compression top dead center, the piston 3 is lowered by motoring. Excessive slope of the heat release rate waveform 6014 due to CI combustion is avoided. Dp / dθ at the time of CI combustion also becomes relatively mild.

  Although dp / dθ can be used as an index representing combustion noise, as described above, SPCCI combustion can reduce dp / dθ, thereby making it possible to avoid combustion noise becoming too large. . Combustion noise can be reduced to an acceptable level or less.

  By the completion of the CI combustion, the SPCCI combustion is ended. CI combustion has a shorter combustion period than SI combustion. SPCCI combustion has an earlier combustion finish timing than SI combustion. In other words, SPCCI combustion can bring the combustion end timing during the expansion stroke closer to the compression top dead center. SPCCI combustion is more advantageous for improving the fuel consumption performance of the engine 1 than SI combustion.

  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 engine 1 is operating in the low load range (1) -1. Specifically, by providing a positive overlap period in which both the intake valve 21 and the exhaust valve 22 are opened near the exhaust top dead center, the combustion chamber 17 is exhausted to the intake port 18 and the exhaust port 19. A portion of the exhaust gas is reintroduced into the combustion chamber 17. Since hot burned gas is introduced into the combustion chamber 17, the temperature in the combustion chamber 17 can be increased, which is advantageous for stabilization of SPCCI combustion. A negative overlap period may be provided in which both the intake valve 21 and the exhaust valve 22 are closed.

  In addition, when the engine 1 is operating in the low load range (1) -1, a strong swirl flow is formed in the combustion chamber 17. The swirl flow is strong at the outer peripheral portion of the combustion chamber 17 and weak at the central portion. The swirl control valve 56 is a predetermined opening on the fully closed or closed side. As described above, since the intake port 18 is a tumble port, an oblique swirl flow having a tumble component and a swirl component is formed in the combustion chamber 17.

  When the engine 1 operates in the low load range (1) -1, the swirl ratio is 4 or more. Here, when the swirl ratio is defined, the “swirl ratio” is a value obtained by dividing the 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 determined based on measurements using the rig test apparatus shown in FIG. That is, in the 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 supply device (not shown) while the cylinder 36 is installed on the cylinder head 13 And an impulse meter 38 having a honeycomb rotor 37 at its upper end. The lower surface of the impulse meter 38 is positioned at 1.75 D (where D is a cylinder bore diameter) from the mating surface of the cylinder head 13 and the cylinder block. The torque acting on the honeycomb rotor 37 can be measured by the impulse meter 38 by the swirl (see the arrow in FIG. 8) generated in the cylinder 36 according to the intake supply, and the intake flow lateral angular velocity can be determined based thereon. .

  FIG. 9 shows the relationship between the opening degree of the swirl control valve 56 in the engine 1 and the swirl ratio. FIG. 9 shows the opening degree of the swirl control valve 56 by the opening ratio with respect to the full open cross section of the secondary passage 402. When the swirl control valve 56 is fully closed, the opening ratio of the secondary passage 402 is 0%, and when the opening degree of the swirl control valve 56 is large, the opening ratio of the secondary passage 402 is 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. 9, when the swirl control valve 56 is fully closed, the swirl ratio becomes approximately 6 when the swirl control valve 56 is fully closed. When the engine 1 operates in the low load range (1) -1, the swirl ratio may be 4 or more and 6 or less. The opening degree of the swirl control valve 56 may be adjusted in the range where the opening ratio is 0 to 15%.

  When the engine 1 operates in the low load range (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. That is, 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 mixture in the entire combustion chamber 17 is 30 or more. By doing this, the generation of RawNOx can be suppressed, and the exhaust gas performance can be improved.

  When the engine 1 operates in the low load range (1) -1, the 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. A central portion in the combustion chamber 17 may be defined as a portion where the swirl flow is weak, and an outer peripheral portion may be defined as a portion where the swirl flow is strong.

  The fuel concentration of the air-fuel mixture in the central portion is higher than the fuel concentration in the outer peripheral portion. Specifically, the A / F of the mixture at the central portion is 20 or more and 30 or less, and the A / F of the mixture at the outer peripheral portion is 35 or more. The value of the air-fuel ratio is the value of the air-fuel ratio at the time of ignition, and is the same in the following description.

  When the engine 1 operates in the low load range (1) -1, the injector 6 injects fuel into the combustion chamber 17 multiple times during the compression stroke (reference numerals 6011 and 6012). As described above, the air-fuel mixture is stratified in the central portion and the outer peripheral portion of the combustion chamber 17 by the multiple fuel injections and the swirl flow in the combustion chamber 17.

  After the end of the fuel injection, the spark plug 25 ignites the mixture in the central portion of the combustion chamber 17 at a predetermined timing before compression top dead center (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 SI combustion, CI combustion starts at an appropriate timing. In SPCCI combustion, the controllability of CI combustion is improved. As a result, when the engine 1 operates in the low load range (1) -1, both suppression of the generation of combustion noise and improvement of the fuel efficiency performance by shortening the combustion period are compatible.

  As described above, in the low load range (1) -1, the engine 1 leans the air-fuel mixture to the stoichiometric air fuel ratio to perform SPCCI combustion, so the low load range (1) -1 is the "SPCCI lean range" It can be called.

(Medium load area (1) -2)
Even when the engine 1 is operating in the medium load range (1) -2, the engine 1 performs SPCCI combustion as in the low load range (1) -1.

  Reference numeral 602 in FIG. 7 indicates the fuel injection timing (reference numerals 6021 and 6022) and the ignition timing (reference numeral 6023) when the engine 1 is operating in the operation state 602 in the medium load range (1) -2. A waveform (symbol 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 range (1) -2. Specifically, as in the low load region (1) -1, by providing a positive overlap period in which both the intake valve 21 and the exhaust valve 22 are opened in the vicinity of the exhaust top dead center, A part of the exhaust gas exhausted from the inside to the intake port 18 and the exhaust port 19 is reintroduced into the combustion chamber 17. That is, the internal EGR gas is introduced into the combustion chamber 17. Further, in the middle load region (1) -2, the exhaust gas cooled by the EGR cooler 53 is introduced into the combustion chamber 17 through the EGR passage 52. That is, the external EGR gas whose temperature is lower than that of the internal EGR gas is introduced into the combustion chamber 17. In the medium load area (1) -2, the internal EGR gas and / or the external EGR gas are introduced into the combustion chamber 17 to adjust the temperature in the combustion chamber 17 to be appropriate.

  Also, even when the engine 1 operates in the medium load range (1) -2, a strong swirl flow having a swirl ratio of 4 or more in the combustion chamber 17 as in the low load range (1) -1. Is formed. The swirl control valve 56 is a predetermined opening on the fully closed or closed side. Since the turbulent energy in the combustion chamber 17 is increased by intensifying the swirl flow, the SI combustion flame propagates rapidly when the engine 1 is operated in the medium load range (1) -2, and the SI combustion is performed. Is stabilized. The stability of the SI combustion improves the controllability of the CI combustion. By optimizing the timing of the CI combustion in the SPCCI combustion, it is possible to suppress the generation of combustion noise and to improve the fuel consumption performance. In addition, variation in torque between cycles can be suppressed.

  When the engine 1 operates in the medium load range (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 mixture may be made to fall within the purification window of the three-way catalyst. Therefore, the excess air ratio λ of the mixture may be 1.0 ± 0.2.

  When the engine 1 operates in the medium load range (1) -2, the injector 6 performs fuel injection (code 6021) during the intake stroke and fuel injection (code 6022) during the compression stroke. 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 reduced by the latent heat of vaporization of the fuel. Pre-ignition of the mixture including the fuel injected by the first injection 6021 can be prevented. In the middle load range (1) -2, in particular, when the engine is in a low load operation state, the second injection 6022 can be omitted.

  By the injector 6 performing the first injection 6021 and the second injection 6022, substantially homogeneous mixing in which the excess air ratio λ becomes 1.0 ± 0.2 as a whole in the combustion chamber 17 My mind is formed. Since the mixture is substantially homogeneous, it is possible to improve the fuel efficiency by reducing the unburned loss, and to improve the exhaust gas performance by avoiding the generation of smoke. The excess air ratio λ is preferably 1.0 to 1.2.

  The fuel-air mixture burns by flame propagation by the spark plug 25 igniting the air-fuel mixture at a predetermined timing before compression top dead center (reference numeral 6023). After the start of combustion by flame propagation, the unburned mixture is self-ignited to burn CI.

  Therefore, in the medium load range (1) -2, the engine 1 performs SPCCI combustion with the air-fuel mixture at the stoichiometric air fuel ratio, so the medium load range (1) -2 can be called "SPCCI λ = 1 range" .

  Here, as shown in FIG. 6, the turbocharger 44 is turned off in a part of the low load area (1) -1 and a part of the medium load area (1) -2. See C OFF). Specifically, in the low rotation region in the low load region (1) -1, the supercharger 44 is turned off. In the high rotation side region in the low load region (1) -1, the supercharger 44 is turned on in order to secure the necessary intake charge 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 supercharger 44 is turned on in order to secure the necessary intake charge amount corresponding to the increase of the engine speed, and it is necessary to cope with the increase in the rotational speed of the engine 1 in the high rotation side region. The turbocharger 44 is turned on to ensure the intake charge.

  In each of the high load / medium speed rotation area (2), the high load / low speed rotation area (3), and the high speed rotation area (4), the turbocharger 44 is turned on over the entire area (S / See C ON).

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

  The reference numeral 603 in FIG. 7 indicates the fuel injection timing (reference numerals 6031 and 6032) and the ignition timing (reference numeral 6033) when the engine 1 is operated in the low load operating condition 603 in the high load medium rotation region (2). , And a combustion waveform (symbol 6034) are shown. Further, reference numeral 604 indicates a fuel injection timing (reference numeral 6041) and an ignition timing (reference numeral 6042) when the engine 1 is operated in the high load side rotation state (2) in the high rotation side operation state 604; The combustion waveform (symbol 6043) 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 medium rotation region (2). The engine 1 reduces the amount of EGR gas as the load increases. At full load, the EGR gas may be zero.

  Also, even when the engine 1 is operated in the high load / medium rotation region (2), as in the low load region (1) -1, a strong swirl flow with a swirl ratio of 4 or more in the combustion chamber 17 Is formed. The swirl control valve 56 is a predetermined opening on the fully closed or closed side.

  When the engine 1 is operated in the high load medium rotation region (2), the air fuel ratio (A / F) of the 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 is operated in the operating state 603 in the high load medium rotation region (2), the injector 6 performs the pre-stage injection 6031 in the intake stroke and performs the post-stage injection 6032 in the compression stroke. The pre-injection may start, for example, in the first half of the intake stroke, and the post-injection may, for example, be performed at the end of the compression stroke. The first half of the intake stroke may be the first half of the intake stroke divided into the first and second halves. Specifically, the pre-injection may start fuel injection, for example, at 280 ° CA before compression top dead center.

  When the start of injection of the pre-injection 6031 is in the first half of the intake stroke, although not shown, the fuel spray hits the opening edge of the cavity 31 so that part of the fuel is in the squish area 171 of the combustion chamber 17 The area outside of 31 (see FIG. 2) enters the area within cavity 31 with the remaining fuel. The swirl flow is strong at the outer peripheral portion of the combustion chamber 17 and weak at the central portion. Therefore, part of the fuel entering the squish area 171 enters the swirl flow, and the remaining fuel entering the area within the cavity 31 enters the swirl flow. The fuel that has entered the swirl flow remains in the swirl flow during the intake stroke and the compression stroke, and forms an air-fuel mixture for CI combustion at the outer peripheral portion of the combustion chamber 17. The fuel entering the swirl flow also stays inside the swirl flow during the intake stroke and the compression stroke, and forms a mixture for SI combustion in the central portion of the combustion chamber 17.

  When the engine 1 is operated in the high load medium rotation region (2), the fuel concentration of the mixture at the outer peripheral portion of the combustion chamber 17 is higher than the fuel concentration of the mixture at the central portion and the fuel of the mixture at the outer peripheral portion Make the amount larger than the fuel amount of the mixture in the central part. Specifically, the air-fuel ratio at the central portion where the spark plug 25 is disposed has an excess air ratio λ of 1 or less, and the mixture at the outer peripheral portion has an excess air ratio λ of less than 1. The air-fuel ratio (A / F) of the air-fuel mixture in the central portion may be, for example, 13 or more and the theoretical air-fuel ratio (14.7) or less. In addition, the air-fuel ratio of the air-fuel mixture in the outer peripheral portion may be, for example, 11 or more, the stoichiometric air-fuel ratio or less, or 11 or more and 12 or less. The temperature of the outer peripheral portion of the combustion chamber 17 is lowered by the latent heat of vaporization of the fuel because the amount of fuel in the air-fuel mixture increases. The air-fuel ratio of the entire mixture in the combustion chamber 17 may be 12.5 or more, the stoichiometric air-fuel ratio or less, or 12.5 or more and 13 or less.

  The final stage of the compression process may be the final stage when the compression process is divided into three equal parts: initial, middle, and final. The post-stage injection 6032 performed at the end of the compression stroke may start fuel injection, for example, 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 reduced by the latent heat of vaporization of the fuel. The low-temperature oxidation reaction proceeds during the compression stroke, and the fuel injected by the pre-injection 6031 shifts to the high-temperature oxidation reaction before top dead center, but the post-injection 6032 is performed just before the top dead center By lowering the temperature in the room, 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. The ratio of the injection amount of the pre-stage injection and the injection amount of the post-stage injection may be, for example, 95: 5.

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

  When the engine 1 operates in the operating state 604 in the high load / medium rotation region (2), the injector 6 starts fuel injection in the intake stroke (reference numeral 6041).

  The injection 6041 starting in the intake stroke may start fuel injection in the first half of the intake stroke, as described above. Specifically, the injection 6041 may start fuel injection at 280 ° CA before top dead center. The end of injection 6041 may be in the compression stroke beyond the intake stroke. By setting the start of the injection 6041 to the first half of the intake stroke, as described above, the mixture for CI combustion is formed in the outer peripheral portion of the combustion chamber 17 and the mixture for SI combustion in the central portion of the combustion chamber 17 You can form your mind. The air-fuel mixture in the central portion where the spark plug 25 is disposed preferably has an air excess ratio λ of 1 or less, and the air-fuel ratio of the outer peripheral portion has an air excess ratio λ of 1 or less, preferably 1 Less than. The air-fuel ratio (A / F) of the air-fuel mixture in the central portion 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 in the central portion may be leaner than the stoichiometric air-fuel ratio. In addition, the air-fuel ratio of the air-fuel mixture in the outer peripheral portion may be, for example, 11 or more, the stoichiometric air-fuel ratio or less, or 11 or more and 12 or less. The air-fuel ratio of the entire mixture in the combustion chamber 17 may be 12.5 or more, the stoichiometric air-fuel ratio or less, or 12.5 or more and 13 or less.

  As the rotational speed of the engine 1 increases, the time for which the fuel injected by the injection 6041 reacts decreases. Therefore, it is possible to omit the post-stage injection for suppressing the reaction of the air-fuel mixture.

  The spark plug 25 ignites the mixture in the central portion of the combustion chamber 17 near the compression top dead center (reference numeral 6042). The spark plug 25 ignites, for example, after compression top dead center.

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

  By stratifiing the mixture in the combustion chamber 17 and generating a strong swirl flow in the combustion chamber 17, SI combustion can be sufficiently performed by the start of CI combustion. As a result, the generation of combustion noise can be suppressed, and the generation of NOx can also be suppressed without the combustion temperature becoming too high. In addition, variation in torque between cycles can be suppressed.

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

  As described above, in the high load medium rotation region (2), the engine 1 performs the SPCCI combustion by making the air-fuel mixture richer than the theoretical air fuel ratio or the theoretical air fuel ratio, so the high load middle rotation region (2) It can be referred to as “SPCCI λ ≦ 1 region”.

(High load low rotation area (3))
When the rotational speed of the engine 1 is low, the time required for the crank angle to change by 1 ° becomes long. The reaction of the fuel injected into the combustion chamber 17 may proceed too much, which may cause premature ignition even if SPCCI combustion is performed. Therefore, when the engine 1 is operating in the high load medium rotation region (3), the engine 1 performs not the SPCCI combustion but the SI combustion.

  The reference numeral 605 in FIG. 7 indicates the fuel injection timing (reference numerals 6051 and 6052) and the ignition timing (reference numeral 6053) when the engine 1 is operating in the operation state 605 in the high load medium rotation region (3) and the combustion A waveform (symbol 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 medium rotation region (3). The engine 1 reduces the amount of EGR gas as the load increases. At full load, the EGR gas may be zero.

  When the engine 1 is operating in the high load low rotation range (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 mixture may be made to fall within the purification window of the three-way catalyst. Therefore, the excess air ratio λ of the mixture may be 1.0 ± 0.2. By setting the air-fuel ratio of the air-fuel mixture to the stoichiometric air-fuel ratio, the fuel efficiency performance is improved in the high load low rotation region (3). When the engine 1 operates in the high load low rotation range (3), the fuel concentration of the entire mixture in the combustion chamber 17 is 1 or less at the excess air ratio λ and in the high load middle rotation range (2) It may be made larger than excess air ratio λ in.

When the engine 1 operates in the high load low rotation range (3), the injector 6 injects fuel in the intake stroke (6051), and a period from the end of the compression stroke to the beginning of the expansion stroke (hereinafter referred to as “period”). Fuel is injected into the combustion chamber 17 at a timing (referred to as a retard period) (reference numeral 6052). The initial stage of the expansion stroke may be the initial stage when the expansion stroke is divided into three equal parts: initial, middle and final.
By injecting the fuel during the intake stroke (reference numeral 6051), the formation time of the air-fuel mixture can be sufficiently secured. Further, by injecting the fuel within the retard period (reference numeral 6052), the gas flow in the combustion chamber 17 can be strengthened immediately before the ignition. The fuel pressure is set to a high fuel pressure of 30 MPa or more. By increasing the fuel pressure, the fuel injection period and the mixture formation period can be shortened, and the gas flow in the combustion chamber 17 can be made stronger. The upper limit of the fuel pressure may be, for example, 120 MPa.

  After injecting the fuel, the spark plug 25 ignites the air-fuel mixture at a timing near the compression top dead center (reference numeral 6053). The spark plug 25 may perform ignition after compression top dead center. The mixture performs SI combustion in the expansion stroke. Since SI combustion starts in the expansion stroke, CI combustion does not start.

  The injector 6 may retard the timing of fuel injection as the rotational speed of the engine 1 decreases, in order to avoid pre-ignition. Fuel injection within the retard period may end in the expansion stroke.

  When the engine 1 operates in the high load low rotation region (3), the time from the start of fuel injection to the ignition within the retard period is short. In order to improve the ignitability of the mixture and to stabilize SI combustion, the fuel needs to be transported to the vicinity of the spark plug 25 promptly.

  When the injector 6 injects fuel in the period from the end of the compression stroke to the beginning of the expansion stroke, the fuel spray is mixed with fresh air because the piston 3 is located near the compression top dead center. It flows downward along the portion 311 and radially outward from the center of the combustion chamber 17 along the bottom and circumferential side surfaces of the cavity 31. 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.

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

  The axis of the injection hole of the injector 6 is offset in the circumferential direction with respect to the spark plug 25 so that the spray is illustrated by a two-dot chain line in the upper drawing of FIG. The fuel injected from the injection hole flows in the circumferential direction by the swirl flow in the combustion chamber 17. The swirl flow allows the fuel to be quickly transported near the spark plug 25. The fuel can be vaporized while being transported near the spark plug 25.

  On the other hand, if the swirl flow is too strong, the fuel flows in the circumferential direction and is separated from the vicinity of the spark plug 25. Therefore, when the engine 1 is operated in the high load low rotation area (3), the swirl flow is weaker than that in the high load medium rotation area (2). As a result, the fuel can be transported to the vicinity of the spark plug 25 promptly, so that the ignition performance of the mixture can be improved and the SI combustion can be stabilized.

  In the high load and low rotation region (3), the engine 1 injects the fuel in a retard period from the end of the compression stroke to the beginning of the expansion stroke to perform SI combustion, so the high load and low rotation region (3) It can be called "retarded-SI area".

(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 ° becomes short. Therefore, for example, in the high rotation region in the high load region, it becomes difficult to perform stratification of the mixture in the combustion chamber 17 as described above. When the rotation speed of the engine 1 becomes high, it becomes difficult to perform the above-described SPCCI combustion.

  Therefore, when the engine 1 is operating in the high rotation range (4), the engine 1 performs not the SPCCI combustion but the SI combustion. The high rotation region (4) extends over the entire load direction from low load to high load.

  Reference numeral 606 in FIG. 7 indicates the fuel injection timing (reference numeral 6061) and the ignition timing (reference numeral 6062) when the engine 1 is operating in the high load operation state 606 in the high rotation range (4), and the combustion waveform (Symbol 6063) 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 load, the EGR gas may be zero.

  When operating the engine 1 in the high rotation region (4), the swirl control valve 56 is fully opened. No swirl flow is generated in the combustion chamber 17, and only the tumble flow is generated. By making the swirl control valve 56 fully open, it is possible to increase the filling efficiency in the high rotation area (4) and to reduce the pump loss.

  When the engine 1 operates in the high rotation range (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 mixture may be 1.0 ± 0.2. In the vicinity of the full open load in the high rotation region (4), the air excess ratio λ of the air-fuel mixture may be less than one.

  When the engine 1 operates in the high rotation range (4), the injector 6 starts fuel injection in the intake stroke. The injector 6 injects the fuel at once (reference numeral 6061). By starting fuel injection during the intake stroke, it is possible to form a homogeneous or substantially homogeneous mixture in the combustion chamber 17. Further, when the rotation speed of the engine 1 is high, the vaporization time of the fuel can be secured as long as possible, so that the unburned loss can also be reduced.

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

  Therefore, in the high rotation region (4), the engine 1 starts fuel injection in the intake stroke to perform SI combustion, so the high rotation region (4) can be called "intake-SI region".

(Supercharger control)
As shown in FIG. 6, the supercharger 44 does not drive in a low load and low rotation predetermined operation area (that is, S / C OFF area), and an operation area other than the S / C OFF area (that is, S Drive in the / C ON region). The ECU 10 outputs a control signal to turn on the electromagnetic clutch 45 when the engine 1 is operating in the S / C ON region. As indicated by solid arrows in FIG. 10, the intake air flows into the surge tank 42 after passing through the turbocharger 44 and the intercooler 46. The ECU 10 also outputs a control signal to adjust the opening degree of the air bypass valve 48 of the bypass passage 47 when the engine 1 is operating in the S / C ON region. When the air bypass valve 48 is opened, part of the gas flows back from the surge tank 42 through the bypass passage 47 and upstream of the turbocharger 44, as shown by the broken arrow in FIG. Thereby, the boost pressure of the intake is adjusted.

  The ECU 10 outputs a control signal to turn off the electromagnetic clutch 45 when the engine 1 is operating in the S / C OFF region. The ECU 10 also outputs a control signal to fully open the air bypass valve 48. As indicated by solid arrows in FIG. 11, the intake air flows into the surge tank 42 through the bypass passage 47 without passing through the turbocharger 44 and the intercooler 46.

(Adjustment of temperature in the combustion chamber)
In each region where SPCCI combustion is performed in the operation region map 501 shown in FIG. 6, it is necessary to adjust the inside of the combustion chamber 17 to an appropriate temperature range in order to stably perform the SPCCI combustion. For example, when the outside air temperature is high, the temperature of the intake air introduced into the combustion chamber 17 is high, so the temperature in the combustion chamber 17 may be too high, and the timing of self-ignition may be advanced. In addition, even when the temperature of the engine 1 becomes too high, the temperature in the combustion chamber 17 may become too high, and the timing of self-ignition may be advanced.

  When the engine 1 is operated with the supercharger 44 turned on in the region where SPCCI combustion is performed, intake air passing through the intercooler 46 downstream of the supercharger 44 is introduced into the combustion chamber 17 as shown in FIG. Therefore, even when the outside air temperature is high or the temperature of the engine 1 is high, it is possible to prevent the temperature in the combustion chamber 17 from becoming too high.

  On the other hand, when the engine 1 is operated with the supercharger 44 turned off, as shown in FIG. 11, the intake air does not pass through the intercooler 46, so the intake air can not be cooled using the intercooler 46.

  Therefore, when the engine 1 is operated with the supercharger 44 turned off, the engine 1 changes the effective compression ratio of the engine 1 by changing the closing timing of the intake valve 21 so that the inside of the combustion chamber 17 is changed. Adjust the temperature. Since this control is performed in the S / C OFF region, it is performed in part of the low load region (1) -1 and part of the medium load region (1) -2.

  Here, the change of the valve timing of the intake valve 21 will be described with reference to FIG. As indicated by the valve lift curve 1401, positive overlap periods of the intake valve 21 and the exhaust valve 22 are provided in the low load area (1) -1 and the medium load area (1) -2 (exhaust valve The solid line showing the valve lift of 22 and the broken line showing the valve lift of the intake valve 21). The closing timing of the intake valve 21 is later than the intake bottom dead center, and is in the compression stroke.

  In this state, as shown by a white arrow in FIG. 14, when the valve timing of the intake valve 21 is retarded, the valve lift of the intake valve 21 changes as indicated by a thick solid line, so the intake valve 21 and the exhaust valve As the overlap period of 22 is shortened (see the hatched portion in the same drawing), the closing timing of the intake valve 21 is further delayed during the compression stroke. The intake valve 21 has a so-called late closing.

  Since the timing at which the compression of the combustion chamber 17 starts is delayed by the late closing of the intake valve 21, the effective compression ratio of the engine 1 decreases. As shown by the graph 1402 in FIG. 14, the pressure in the combustion chamber 17 when the piston 3 reaches compression top dead center decreases when the intake valve 21 is closed late (see the white arrow). .

  Further, as the positive overlap period of the intake valve 21 and the exhaust valve 22 becomes shorter, the amount of EGR gas reintroduced into the combustion chamber 17 decreases. As a result, as shown in the graph of reference numeral 1403, the temperature at the start of compression of the combustion chamber 17 decreases.

  When the temperature at the start of compression of the combustion chamber 17 decreases and the effective compression ratio of the engine 1 decreases, as shown in the graph 1403, when the piston 3 reaches compression top dead center The temperature in the combustion chamber 17 (so-called compression end temperature) is lowered and can be within the allowable range.

  In addition, when the engine 1 is operated with the supercharger 44 turned off, the outside temperature is too high, and the temperature in the combustion chamber 17 is appropriately adjusted by changing the effective compression ratio of the engine 1 When adjustment can not be made to the range, cooling using the intercooler 46 is performed by allowing part of the intake air to pass through the turbocharger 44 and the intercooler 46 although the turbocharger 44 is not driven. This control is called pre-rotation control of the supercharger 44 because the supercharger 44 is rotated in the non-driving region of the supercharger 44.

  The prerotation control of the turbocharger 44 will be described with reference to FIGS. 12 and 13. FIG. 12 shows the flow of intake air at the time of prerotation control of the turbocharger 44. The ECU 10 outputs a control signal to turn off the electromagnetic clutch 45 when performing pre-rotation control, and adjusts the opening of the air bypass valve 48 to the closing side. As a result, a pressure difference occurs between the upstream side and the downstream side of the turbocharger 44, so that the rotor of the turbocharger 44 is rotated by the pressure difference. A portion of the intake air passes through the turbocharger 44 and the intercooler 46 and is introduced into the combustion chamber 17 as indicated by a broken line. A portion of the intake air is also introduced into the combustion chamber 17 through the bypass passage 47, as shown by the solid line.

  In the prerotation control of the supercharger 44, by adjusting the opening degree of the air bypass valve 48, the amount of intake air passing through the supercharger 44 and the intercooler 46 and entering the combustion chamber 17 and the bypass passage 47 The ratio to the amount of intake air entering the combustion chamber 17 is changed. Since the amount of intake air cooled by the intercooler 46 changes, the temperature in the combustion chamber 17 changes. FIG. 13 shows the relationship between the flow rate ratio of the intake air passing through the intercooler 46 (horizontal axis) and the temperature in the surge tank 42 (horizontal axis). The flow rate ratio of the intake air passing through the intercooler 46 is related to the opening degree of the air bypass valve (ABV) 48 as described above. When the opening of the air bypass valve 48 is large, the flow ratio of intake air passing through the intercooler 46 decreases, and when the opening of the air bypass valve 48 is small, the flow ratio of intake air passing through the intercooler 46 increases. The temperature in the surge tank 42 corresponds to the temperature of the intake air introduced into the combustion chamber 17. As shown in FIG. 13, the temperature in the surge tank 42 can be lowered by reducing the opening of the air bypass valve 48 and increasing the flow rate ratio of the intake air passing through the intercooler 46. Since the temperature of the intake air introduced into the combustion chamber 17 can be kept within the appropriate range S1, the compression end temperature can be kept within the allowable range.

  Furthermore, when the engine 1 is operated with the supercharger 44 turned off, the temperature inside the combustion chamber 17 is also appropriately set by the pre-rotation of the supercharger 44 described above because the outside air temperature is higher. When the temperature range can not be adjusted, cooling using the intercooler 46 is performed by turning on the turbocharger 44 and circulating a part of the intake air through the turbocharger 44, the intercooler 46 and the bypass passage 47. Do. This control is called circulation control of the turbocharger 44.

  The flow of intake air at the time of circulation control of the supercharger 44 is substantially the same as the flow of intake air at the time of supercharging shown in FIG. That is, the ECU 10 outputs a control signal to turn on the electromagnetic clutch 45 when performing circulation control, and adjusts the opening degree of the air bypass valve 48 to the open side. Thus, the intake air is cooled by passing through the intercooler 46 after passing through the supercharger 44. A part of the cooled intake air is introduced from the surge tank 42 into the combustion chamber 17, and the remaining intake air not introduced into the combustion chamber 17 flows back through the bypass passage 47, and along with the new intake air. , Is reintroduced into the turbocharger 44. Thus, part of the intake air passes through the intercooler 46 a plurality of times, and it is possible to efficiently lower the temperature of the intake air introduced into the combustion chamber 17 when the outside air temperature is extremely high.

(Engine control process)
Next, operation control of the engine 1 executed by the ECU 10 will be described with reference to the flowchart of FIG. This flowchart relates to temperature control in the combustion chamber 17.

  First, in step S1 after the start, the ECU 10 reads the signals of the sensors SW1 to SW16. The ECU 10 determines the operating range of the engine 1 in the subsequent step S2.

  In step S3, the ECU 10 determines whether the operating state of the engine 1 is in the region where the supercharger 44 is turned on. If the determination in step S3 is YES, the process proceeds to step S9. If the determination is NO, the process proceeds to step S4.

  The ECU 10 turns on the electromagnetic clutch 45 of the turbocharger 44 in step S9. Thus, the intake air passes through the turbocharger 44 and the intercooler 46 and is introduced into the combustion chamber 17. Since the intake air is cooled by passing through the intercooler 46, the temperature in the combustion chamber 17 can be kept within an appropriate temperature range even when the outside air temperature is high or the temperature of the engine 1 is high. .

  The ECU 10 turns off the electromagnetic clutch 45 of the turbocharger 44 in step S4. Further, the ECU 10 opens the air bypass valve 48. In this way, the intake air is introduced into the combustion chamber 17 through the bypass passage 47 without passing through the turbocharger 44 and the intercooler 46.

  In step S5, the ECU 10 determines whether the compression end temperature of the combustion chamber 17 exceeds a preset allowable range. The ECU 10 predicts the compression end temperature based on various sensor values. The ECU 10 is an example of a temperature estimation unit. If the determination in step S5 is NO, the process proceeds to step S10. If the determination in step S5 is YES, the process proceeds to step S6.

  In step S6, the ECU 10 determines whether the outside air temperature exceeds a second predetermined temperature set in advance. The second predetermined temperature is a temperature at which the outside air temperature is extremely high. When the determination in step S6 is YES, the process proceeds to step S12, and when the determination is NO, the process proceeds to step S7.

  In step S7, the ECU 10 determines whether the outside air temperature exceeds a first predetermined temperature set in advance. The first predetermined temperature is a temperature lower than the second predetermined temperature. When the determination in step S7 is YES, the process proceeds to step S13, and when the determination is NO, the process proceeds to step S8. That is, in this control process, the outside air temperature is determined in three steps of “very high”, “high”, and “slightly high” in steps S6 and S7. When the outside air temperature is extremely high, the process proceeds to step S12 as described above. If the outside air temperature is high, the process proceeds to step S13. If the outside air temperature is somewhat high, the process proceeds to step S8.

  In step S12, the ECU 10 performs the circulation control of the supercharger 44 described above. That is, in the non-supercharging region, the electromagnetic clutch 45 of the turbocharger 44 is turned on to drive the turbocharger 44, and the opening of the air bypass valve (ABV) 48 is adjusted to the open side. Thereby, as shown in FIG. 10, a part of the intake air is circulated in the supercharger 44, the intercooler 46, and the bypass passage 47 to lower the temperature of the intake air introduced into the combustion chamber 17.

  In step S13, the ECU 10 performs the prerotation control of the supercharger 44 described above. That is, with the electromagnetic clutch 45 of the turbocharger 44 turned off, the opening degree of the air bypass valve 48 is adjusted to the closing side. As a result, as shown in FIG. 12, since the supercharger 44 pre-rotates, the intake air passing through the supercharger 44 and the intercooler 46 can be introduced into the combustion chamber 17. The intake air cooled by the intercooler 46 can be introduced into the combustion chamber 17.

  In step S8, the ECU 10 outputs a control signal to the intake electric motor S-VT 23 so that the valve timing of the intake valve 21 is retarded. By delaying the intake valve 21, the compression end temperature of the combustion chamber 17 is lowered by reducing the effective compression ratio of the engine 1 and shortening the overlap period of the intake valve 21 and the exhaust valve 22 (see FIG. 14). .

  In step S10, the ECU 10 determines whether the compression end temperature is below the allowable range. When NO, the process returns. If yes, the process proceeds to step S11. In step S11, the ECU 10 outputs a control signal to the intake electric motor S-VT 23 so that the valve timing of the intake valve 21 is advanced. By bringing the closing timing of the intake valve 21 closer to the top dead center, the effective compression ratio of the engine 1 becomes higher, and the overlap period of the intake valve 21 and the exhaust valve 22 becomes longer. Contrary to the case where the intake valve 21 is closed late, the compression end temperature of the combustion chamber 17 can be raised, and the compression end temperature can be kept within the allowable range.

(Other embodiments)
The technology disclosed herein is not limited to application to the engine 1 configured as described above. The configuration of the engine 1 can adopt various configurations.

FIG. 16 shows a part of a flowchart showing a reference example of a control process of the engine 1. Steps S14, S15, and S16 in this flowchart relate to switching between the prerotation control of the turbocharger 44 and the late closing of the intake valve 21, and can be replaced with steps S7, S8, and S13 in the flowchart shown in FIG. Can.

  In the flowchart shown in FIG. 15, when the outside air temperature exceeds the first predetermined temperature, prerotation control of the supercharger 44 is performed in step S13, and when the outside air temperature is equal to or less than the first predetermined temperature, intake air is selected in step S8. The valve 21 is closed late. On the other hand, in the flowchart of FIG. 16, the prerotation control of the turbocharger 44 and the late closing of the intake valve 21 are switched depending on whether or not the driver's acceleration request operation is performed.

  Specifically, the ECU 10 determines in step S14 whether or not there is a driver's acceleration request operation. When there is an acceleration request operation, that is, when the driver steps on the accelerator pedal, the process proceeds to step S15, and when there is no acceleration request operation, the process proceeds to step S16.

  When there is an acceleration demand operation, the intake air introduced into the combustion chamber 17 must be increased. In the prerotation control of the turbocharger 44, the turbocharger 44 is rotated by the pressure difference, so that the intake resistance becomes high. Therefore, if the prerotation control of the supercharger 44 is performed at the time of acceleration where increase in intake is required, increase in intake will be delayed. The acceleration response may be reduced.

  Therefore, when there is an acceleration request operation, control is performed to retard the phase of the intake valve 21 in step S15 to lower the temperature in the combustion chamber 17. As a result, the temperature in the combustion chamber 17 can be kept within the allowable range while preventing a decrease in acceleration response.

  When the acceleration request operation is not performed, it is possible to allow the intake resistance to be high. Therefore, in step S16, the prerotation control of the supercharger 44 is performed to lower the temperature in the combustion chamber 17. As a result, the intake air can be cooled efficiently, so the temperature in the combustion chamber 17 can be kept within the allowable range.

  The engine 1 may be provided with a turbocharger instead of the mechanical supercharger 44.

  In addition to using the intake electric motor S-VT 23 that changes the phase of the valve timing of the intake valve 21, the above-described technology also uses a variable valve mechanism that changes the lift amount of the intake valve 21. It can also be realized by using a variable valve mechanism that changes the valve opening angle.

  In addition, the technology disclosed herein can be widely applied not only to engines that perform SPCCI combustion but also to engines that perform self-ignition combustion.

1 Engine 10 ECU (control unit, temperature estimation unit)
17 combustion chamber 23 intake electric motor S-VT (variable valve mechanism)
40 intake passage 44 supercharger 45 electromagnetic clutch (switching unit)
46 Intercooler 47 Bypass passage 48 Air bypass valve (flow control valve)

Claims (5)

An intake passage connected to the combustion chamber of the engine;
A supercharger disposed in the intake passage;
A switching unit that switches between driving and non-driving of the turbocharger;
An intercooler disposed downstream of the supercharger in the intake passage;
A bypass passage connecting the upstream of the turbocharger in the intake passage and the downstream of the intercooler;
A flow control valve disposed in the bypass passage;
A variable valve mechanism configured to change the opening and closing operation of an intake valve interposed between the downstream end of the intake passage and the combustion chamber;
A control unit connected to the switching unit, the flow control valve, and the variable valve mechanism, and outputting a control signal to the switching unit, the flow control valve, and the variable valve mechanism;
A temperature estimation unit for estimating a compression end temperature which is a temperature in the combustion chamber when the piston reaches compression top dead center,
When the control unit operates the engine without driving the supercharger through the switching unit, the control unit opens the flow control valve, and in the engine, the air-fuel mixture in the combustion chamber is burned by compression self-ignition And
The control unit also causes the compression end temperature to be a preset allowable temperature based on the compression end temperature estimated by the temperature estimation unit when the engine is operated with the supercharger not driven. When the outside air temperature exceeds a predetermined temperature, the supercharger is not driven through the switching section, and the flow control valve is operated more than when the compression end temperature is equal to or less than the allowable temperature. Close
The control unit is configured to set the effective compression ratio of the engine when the compression end temperature exceeds the allowable temperature and the outside air temperature is lower than the predetermined temperature, and when the compression end temperature is lower than the allowable temperature. The supercharged compression self-ignition type engine which changes the closing timing of the intake valve through the variable valve mechanism so as to lower also. In the supercharged compression self-ignition engine according to claim 1,
The control unit deactivates the turbocharger when the load on the engine is low,
The variable valve mechanism includes a mechanism for changing the phase of valve timing of the intake valve, and when the load on the engine is low, an overlap period in which both the intake valve and the exhaust valve are open is selected. By providing the EGR gas introduced into the combustion chamber to a predetermined ratio or more,
The control unit also causes the compressor end temperature to exceed the allowable temperature when the engine is operated with the supercharger not driven, and when the outside air temperature is equal to or lower than the predetermined temperature. The valve timing phase of the intake valve through the variable valve mechanism so that the overlap period is shorter and the effective compression ratio of the engine is lower than when the compression end temperature is lower than the allowable temperature. A compression self-ignition engine with a supercharger that retards. In the supercharged compression self-ignition engine according to claim 1 ,
The control unit drives the supercharger through the switching unit when the compression end temperature exceeds the allowable temperature and the outside air temperature exceeds a second predetermined temperature higher than the predetermined temperature. A supercharged compression self-ignition engine that opens the flow control valve. The supercharged compression self-ignition engine according to any one of claims 1 to 3 , wherein
The said engine is a compression self-ignition type | mold engine with a supercharger whose geometric compression ratio is 13-30. The supercharged compression self-ignition engine according to any one of claims 1 to 4 , wherein
The fuel supplied to the combustion chamber is a supercharged compression self-ignition engine including at least gasoline.

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