JP5589959B2 - Spark ignition engine with turbocharger - Google Patents

Description

  The present invention includes a supercharger that pressurizes intake air, an injector that injects fuel into the combustion chamber, an ignition plug that discharges sparks to an air-fuel mixture formed in the combustion chamber based on the fuel injected from the injector, The present invention relates to a spark ignition engine equipped with a supercharger, comprising a supercharger, an injector, and a control means for controlling the operation of each part including an ignition plug.

  Conventionally, as shown in Patent Document 1 below, a spark ignition with a supercharger configured so that HCCI combustion (Homogeneous-Charge Compression Ignition combustion) for burning an air-fuel mixture by self-ignition is performed at least in a partial load region. In the engine, the air excess ratio λ of the air-fuel mixture when the engine is warm is set to 2 or more over the entire engine load.

  In particular, in Patent Document 1, in the high load side region (second HCCI region) in the HCCI combustion execution region, the supercharging amount by the supercharger is increased and the effective compression ratio of the engine based on the closing timing of the intake valve , And by dividing and injecting the fuel multiple times, the lean mixture of λ = 2 is burned by HCCI while the rate of increase in fuel pressure (dp / dθ) is reduced.

JP 2010-236429 A

  When the mixture is HCCI-combusted while reducing the rate of increase in the combustion pressure by reducing the effective compression ratio of the engine or performing split fuel injection as in Patent Document 1 above, combustion noise Appropriate HCCI combustion without any increase or abnormal combustion can be continuously performed up to a relatively high load range.

  However, as described above, in order to continue the HCCI combustion to a considerably high load (for example, in the vicinity of the maximum load), measures such as extremely reducing the effective compression ratio are required, which is not very realistic. Therefore, in Patent Document 1, when the load increases to the vicinity of the maximum load of the engine (SI region), switching to SI combustion in which the air-fuel mixture is burned by flame propagation is triggered by spark ignition of the spark plug.

  Specifically, in the SI combustion execution region, the excess air ratio λ of the air-fuel mixture is maintained at λ = 2, and multiple ignition is performed to cause the spark plug to perform a plurality of consecutive spark ignitions. The In general, under a significantly lean air-fuel ratio such as λ = 2, the flame propagation speed is greatly reduced, and misfire is likely to occur in normal spark ignition, so multiple ignition as described above is necessary.

  As described above, in Patent Document 1, since a significantly lean air-fuel ratio of λ = 2 is maintained up to a high load range of the engine, improvement in thermal efficiency and suppression of NOx generation amount can be expected. However, in order to maintain a lean air-fuel ratio of λ = 2 up to the high load range of the engine, a supercharger with a high supercharging capability capable of pumping a considerably large amount of air commensurate with the injected fuel increased according to the load is required. Not only is it necessary, but the exhaust pressure of the engine (the pressure of the exhaust gas upstream from the turbocharger turbine) increases excessively due to a large amount of supercharging, resulting in an increase in pump loss, resulting in poor thermal efficiency. There is a fear.

  The present invention has been made in view of the circumstances as described above, and it is possible to properly continue combustion with high thermal efficiency and excellent emission properties to a high load range while properly using HCCI combustion and SI combustion. An object is to provide a spark ignition engine with a supercharger.

In order to solve the above problems, the present invention provides a supercharger that pressurizes intake air, an injector that injects fuel into the combustion chamber, and an air-fuel mixture formed in the combustion chamber based on the fuel injected from the injector. And a control unit for controlling the operation of each part including the supercharger, the injector, and the spark plug, and a supercharger in which the geometric compression ratio is set to 14 or more A retarded SI mode for starting the flame propagation combustion by the spark ignition of the spark plug after the compression top dead center at least in a high load region when the engine is warm. On the other hand, in a predetermined load region where the load is lower than the execution region of the retarded SI mode, the air-fuel mixture is burned by self-ignition while supercharging the supercharger. The supercharging HCCI mode is performed. In the supercharging HCCI mode, the excess air ratio λ of the air-fuel mixture is set to λ ≧ 2 by introducing a large amount of air into the combustion chamber by supercharging. Control is performed to burn a lean air-fuel mixture with λ ≧ 2 from the vicinity of compression top dead center by self-ignition. In the retarded SI mode, the excess air ratio λ of the air-fuel mixture is set to λ = 1, and from the injector the fuel injection executed in the compression stroke later due to higher injection pressure 20 MPa, by further performing the spark ignition by the spark plug Rise Zhang stroke Initial, a mixture of the compression top dead center after waiting more than a predetermined time period Control that causes rapid combustion by flame propagation is executed (Claim 1).

  According to the present invention, a mixture of λ ≧ 2 based on a large amount of air supercharged by a supercharger is burned by self-ignition (HCCI combustion) in a predetermined load range when the engine is warm. Since the combustion temperature is sufficiently low, the NOx amount (raw NOx amount) generated by combustion can be greatly reduced, and the thermal efficiency can be effectively improved.

  On the other hand, on the higher load side than the HCCI combustion execution region when the engine is warm, high-pressure injection of 20 MPa or more is executed at a later timing after the later stage of the compression stroke, and λ = 1 based on the mixture When spark ignition is performed on the air-fuel mixture, the air-fuel mixture is rapidly burned by flame propagation (SI after the compression top dead center has passed (that is, after the in-cylinder temperature and pressure have dropped to some extent). Combustion). Therefore, unlike conventional SI combustion in which fuel is injected at an early timing such as the intake stroke, it is possible to realize combustion with excellent thermal efficiency with a short combustion period while reliably avoiding abnormal combustion such as pre-ignition and knocking. Can do. In particular, since the excess air ratio λ of the air-fuel mixture is set to λ = 1 (theoretical air-fuel ratio), unlike the case where a lean air-fuel ratio of λ ≧ 2 is maintained, the supercharging amount of intake air is extremely increased. There is no need to increase the pump loss due to excessive increase in the exhaust pressure of the engine and the deterioration of the fuel consumption based on it. Further, if combustion is performed at λ = 1, NOx can be sufficiently purified only by the three-way catalyst, and the emission property is also ensured satisfactorily.

  In the present invention, preferably, the injector is a multi-hole injector having a plurality of nozzle holes (Claim 2).

  According to this configuration, in the retarded SI mode, the turbulent energy generated based on the fuel injection of the injector is further increased and the vaporization of the fuel is further promoted, so that the combustion period is further shortened. Thermal efficiency can be improved.

  In the present invention, preferably, the engine is equipped with an external EGR device that recirculates exhaust gas discharged from the combustion chamber into the exhaust passage to the intake passage. In the supercharged HCCI mode, recirculation of exhaust gas by the external EGR device. The operation is executed (claim 3).

  According to this configuration, during operation in the supercharged HCCI mode, NOx is generated by further reducing the combustion temperature of the air-fuel mixture while appropriately securing the supercharging amount for realizing a lean air-fuel ratio of λ ≧ 2. There is an advantage that the amount can be effectively reduced.

  In the present invention, preferably, the control means is a NA-HCCI mode in which an air-fuel mixture is burned by self-ignition without performing supercharging by the supercharger on a lower load side than the execution region of the supercharging HCCI mode. In the NA-HCCI mode, internal EGR is performed in which high-temperature exhaust gas generated by combustion remains in the combustion chamber (claim 4).

  According to this configuration, in the operating region where the engine load is low and the ignitability of the air-fuel mixture is poor, the air-fuel mixture can be reliably combusted by self-ignition by making full use of the internal EGR to increase the temperature of the combustion chamber. .

  In the present invention, preferably, the supercharger includes a turbine disposed in an exhaust passage of the engine and a compressor disposed in an intake passage of the engine and driven by rotation of the turbine. Between the passage and the exhaust passage, a first external EGR device that recirculates the exhaust gas branched from the exhaust passage downstream of the turbine to the intake passage, and the exhaust gas branched from the exhaust passage upstream of the turbine. A second external EGR device that recirculates to the intake passage is provided, and exhaust gas recirculation operation by the first external EGR device is executed in a part of the low rotation side in the execution region of the retard SI mode, In the remaining area in the execution area of the retarded SI mode, the exhaust gas recirculation operation is performed by the second external EGR device.

  As described above, in the operation region where the retarded SI mode is performed, the exhaust gas after passing through the turbine is recirculated to the intake passage using the first external EGR device in the operation region where the rotational speed is low and the exhaust gas flow rate is small. On the other hand, when the exhaust gas before passing through the turbine is recirculated to the intake passage using the second external EGR device in the operation region where the rotational speed is high and the exhaust gas flow rate is large, There is an advantage that both supercharging performance can be secured and pump loss on the high rotation side can be reduced.

  In the present invention, preferably, the engine is equipped with a variable mechanism that variably sets at least the closing timing of the intake valve, and the control means supercharges the supercharger in a medium to high load range when the engine is cold. The supercharging SI mode in which the air-fuel mixture is combusted by spark ignition is performed while the variable mechanism is driven and the variable mechanism is driven at least at a part of the high load side in the supercharging SI mode execution region. The closing time is set earlier than the bottom dead center (claim 6).

  According to this configuration, even when the engine is cold, in situations where abnormal combustion (pre-ignition or knocking) is likely to occur due to high load, the intake valve closing timing is made earlier than the intake bottom dead center. It is possible to prevent the abnormal combustion by reducing the effective compression ratio of the engine. In addition, unlike the case where the closing timing of the intake valve is made later than the intake bottom dead center due to a decrease in the effective compression ratio, the temperature of the combustion chamber is increased due to the air blown back to the intake port being introduced again into the combustion chamber. Can be avoided, and the effect of preventing abnormal combustion can be further enhanced. As a result, the reduction range of the effective compression ratio can be kept low, and the sacrifice of engine output and fuel consumption can be minimized.

  As described above, according to the spark ignition engine with a supercharger according to the present invention, while using HCCI combustion and SI combustion properly, combustion with high thermal efficiency and excellent emission characteristics is properly continued to a high load range. be able to.

1 is a diagram illustrating an overall configuration of a spark ignition engine with a supercharger according to an embodiment of the present invention. It is a figure which shows the structure of the engine main body of the said engine. It is a block diagram which shows the control system of the said engine. It is a figure which shows an example of the area | region determination map used at the time of engine warm. It is a figure which shows an example of the area | region determination map used when an engine is cold. It is a time chart for demonstrating the control content of the NA-HCCI mode performed in the 1st driving | operation area | region (A1) of FIG. FIG. 5 is a time chart for explaining the control contents of a supercharging HCCI mode executed in a second operation region (A2) of FIG. 4. FIG. 5 is a time chart for explaining the control contents of first and second retarded SI modes executed in the third and fourth operation regions (A3, A4) of FIG. 4. It is a figure for demonstrating the control example (the 1) of the intake / exhaust control using a 1st EGR valve, a 2nd EGR valve, and a waste gate valve. It is a figure for demonstrating the control example (the 2) of the intake / exhaust control using a 1st EGR valve, a 2nd EGR valve, and a waste gate valve. It is a figure for demonstrating the control example (the 3) of the intake / exhaust control using a 1st EGR valve, a 2nd EGR valve, and a waste gate valve. It is a figure for demonstrating the characteristic of SI combustion implement | achieved by the said 1st, 2nd retarded SI mode compared with the conventional SI combustion. It is a figure for demonstrating how the high pressure of fuel injection and retarding of injection timing contribute to prevention of abnormal combustion. (A)-(f) is a figure which shows the change of various control parameters when only a load changes in the low speed area of an engine. (A)-(f) is a figure which shows the change of various control parameters when only a load changes in the high speed area of an engine. It is a figure for demonstrating what kind of timing the closing timing of an intake valve is set at least in the high load area at the time of engine cold.

(1) Overall Configuration of Engine FIG. 1 is a diagram showing an overall configuration of an engine according to an embodiment of the present invention, and FIG. 2 is a cross-sectional view showing a specific configuration of the engine body 1. The engine shown in these drawings is a reciprocating piston type multi-cylinder gasoline engine mounted on a vehicle as a power source for driving driving. An engine main body 1 of this engine is slidable back and forth in a cylinder block 2 having a plurality of cylinders 2A to 2D arranged in a uniaxial direction, a cylinder head 3 provided on the upper surface of the cylinder block 2, and each cylinder 2A to 2D. It has an inserted piston 4. In addition, the fuel supplied to the engine body 1 may be anything that contains gasoline as a main component, and the contents may be all gasoline, or gasoline containing ethanol (ethyl alcohol) or the like. But you can.

  The piston 4 is connected to a crankshaft 13 via a connecting rod 12 so that the crankshaft 13 rotates about a central axis in accordance with the reciprocating motion of the piston 4.

  A combustion chamber 5 is formed above the piston 4, an intake port 6 and an exhaust port 7 are opened in the combustion chamber 5, and an intake valve 8 and an exhaust valve 9 that open and close the ports 6 and 7 are connected to the cylinder head 3. Are provided respectively. The engine shown in the figure is a so-called double overhead camshaft (DOHC) engine, in which two intake ports 6 and two exhaust ports 7 are provided for each cylinder, and two intake valves 8 and two exhaust valves 9 are provided. It is provided one by one.

  The “combustion chamber”, in a narrow sense, refers to a space formed above the piston 4 when the piston 4 rises to the top dead center. The space formed above the piston 4 regardless of the vertical position of the piston 4 (a combustion chamber in a broad sense).

  Here, the geometric compression ratio of the engine body 1 of the present embodiment is set to a relatively high value of 14 or more for the purpose of improving the theoretical thermal efficiency, stabilizing HCCI combustion (compression self-ignition combustion), which will be described later, or the like. Is set. Since the upper limit value of the geometric compression ratio is considered to be about 20 from a practical viewpoint, the geometric compression ratio of the engine body 1 is set to an appropriate value in the range of 14 to 20. Is set.

  As shown in FIG. 2, the intake valve 8 and the exhaust valve 9 rotate the crankshaft 13 by valve mechanisms 15 and 16 each including a pair of camshafts (not shown) disposed in the cylinder head 3. It is driven to open and close in conjunction with.

  A variable mechanism 17 for changing the lift characteristic of the intake valve 8 is incorporated in the valve mechanism 15 for the intake valve 8. The variable mechanism 17 may be any mechanism that can change at least the closing timing of the intake valve 8. However, it is more preferable that the variable mechanism 17 can be operated to change only the closing timing of the intake valve 8 and not to change the opening timing. . As the variable mechanism 17 having such a configuration, for example, a VVT (variable valve timing mechanism) that changes the operation timing (phase angle) of the intake valve 8 and a VVL that changes the lift amount of the intake valve 8 are used. A combination of (Variable Valve Lift Mechanism) is conceivable. With this type of variable mechanism, not only the closing timing of the intake valve 8 but also the valve opening period (that is, the period from when the intake valve 8 starts to be closed) can be changed. It is possible to perform an operation of changing (without changing the opening time).

  The valve mechanism 16 for the exhaust valve 9 incorporates a switching mechanism 18 that enables or disables the function of depressing the exhaust valve 9 during the intake stroke. That is, the switching mechanism 18 allows the exhaust valve 9 to be opened not only in the exhaust stroke but also in the intake stroke, and opens the exhaust valve 9 during the intake stroke (so-called exhaust valve 9 is opened twice). It has a function to switch between executing and stopping.

  When the exhaust valve 9 is opened during the intake stroke by the action of the switching mechanism 18, the high-temperature exhaust gas once discharged to the exhaust port 7 flows back into the combustion chamber 5 and remains in the combustion chamber 5 until the next exhaust stroke. . In the following, the exhaust gas residual operation by opening the exhaust valve 9 twice (opening during the intake stroke) is distinguished from the exhaust gas recirculation operation (external EGR) by the external EGR devices 39 and 44 described later. , Referred to as internal EGR.

  The variable mechanism 17 and the switching mechanism 18 may be appropriately selected from various conventional structures having similar functions.

  The cylinder head 3 of the engine body 1 is provided with one set of injectors 10 and spark plugs 11 for each of the cylinders 2A to 2D.

  The injector 10 is provided so as to face the combustion chamber 5 of each of the cylinders 2A to 2D from the side of the intake side. The injector 10 is a multi-hole type injector having a plurality of nozzle holes at the tip thereof. In the present embodiment, a total of twelve nozzle holes are provided at the tip of the injector 10.

  Moreover, the pressure of the fuel (fuel mainly composed of gasoline) injected from each injection hole of the injector 10 is set to a high value of 20 MPa or more. The injection pressure from the injector 10 is more preferably set to a value higher than 20 MPa (for example, 30 to 40 MPa), but from a practical viewpoint, the upper limit of the injection pressure is considered to be about 120 MPa. . From this, the injection pressure from the injector 10 is set to an appropriate value in the range of 20 MPa to 120 MPa.

  In order to realize the high injection pressure as described above, for example, a fuel pump is provided between a fuel supply pipe connected to the injector 10 of each cylinder 2A to 2D and a fuel pump provided on the upstream side thereof. It is conceivable to adopt a so-called common rail type fuel supply system provided with a common rail for accumulating and storing fuel pumped from the vehicle.

  The spark plug 11 is provided so as to face the combustion chamber 5 of each cylinder 2A to 2D from above. Specifically, the spark plug 11 has an electrode exposed to the combustion chamber 5 at the tip, and discharges a spark from the electrode in response to power supply from an ignition circuit (not shown).

  As shown in FIG. 2, a water jacket (not shown) through which cooling water flows is provided inside the cylinder block 2 and the cylinder head 3 of the engine body 1, and the temperature of the cooling water in the water jacket. The cylinder block 2 is provided with a water temperature sensor SW1 for detecting the above.

  The cylinder block 2 is provided with a crank angle sensor SW2. The crank angle sensor SW2 outputs a pulse signal according to the rotation of a crank plate (not shown) that rotates integrally with the crankshaft 13, and based on this pulse signal, the rotation angle (crank angle) of the crankshaft 13 is output. ) And rotation speed (engine rotation speed) are detected.

  The cylinder head 3 is provided with a cam angle sensor SW3 for detecting the camshaft angle in the valve mechanism 15. The cam angle sensor SW3 is a pulse signal for determining a cylinder (determining whether each cylinder is in an intake stroke, compression stroke, expansion stroke, or exhaust stroke) according to passage of teeth of a signal plate that rotates integrally with the camshaft. Is output.

  An intake passage 20 and an exhaust passage 30 are connected to the intake port 6 and the exhaust port 7 of the engine body 1, respectively. That is, outside intake air (fresh air) is supplied to the combustion chamber 5 through the intake passage 20, and exhaust gas (burned gas) generated in the combustion chamber 5 is discharged to the outside through the exhaust passage 30. It has become so.

  As shown in FIG. 1, the intake passage 20 has a common pipe portion 23 formed of a single pipe and a downstream end of the common pipe portion 23 (a downstream end in the intake air flow direction indicated by an arrow in the drawing). It has a connected surge tank 22 and a branch pipe portion 21 that is branched for each of the cylinders 2A to 2D and connects the surge tank 22 and the intake port 6 of each cylinder 2A to 2D.

  The exhaust passage 30 is provided with a common pipe portion 32 made of a single pipe and branched for each of the cylinders 2A to 2D, and an upstream end of the common pipe portion 32 (flow of exhaust gas indicated by an arrow in the drawing). And a branch pipe portion 31 that connects the exhaust ports 7 of the cylinders 2A to 2D.

  The engine of the present embodiment is equipped with an exhaust turbo-type supercharger 35 and first and second external EGR devices 39 and 44 that recirculate the exhaust gas discharged to the exhaust passage 30 to the intake passage 20. ing.

  The supercharger 35 connects the turbine 36 disposed in the common pipe portion 32 of the exhaust passage 30, the compressor 37 provided in the common pipe portion 23 of the intake passage 20, and the turbine 36 and the compressor 37. And a connecting shaft 38. When the exhaust gas flows into the turbine 36 through the exhaust passage 30 during the operation of the engine, the turbine 36 is rotated by the energy of the exhaust gas, and the compressor 37 is rotated at a high speed in conjunction with this, thereby the intake passage. Air passing through 20 (intake air) is fed into the combustion chamber 5 while being pressurized.

  The common pipe portion 32 of the exhaust passage 30 is provided with a bypass pipe 48 for bypassing the turbine 36 of the supercharger 35. The bypass pipe 48 is provided with a waste gate valve 49, and the exhaust gas flow path is switched according to the opening and closing of the valve 49 so that the supercharger 35 can be switched on / off. It has become.

  That is, when the waste gate valve 49 is closed, exhaust gas flows into the turbine 36, whereby the compressor 37 is driven to rotate and the intake air is pressurized, while the waste gate valve 49 is opened. When the exhaust gas mainly passes through the bypass pipe 48 (bypasses the turbine 36), the pressurization of the intake air by the compressor 37 is substantially stopped. It is also possible to adjust the amount of work (supercharging capability) of the compressor 37 by setting the wastegate valve 49 to an intermediate opening less than 100% (fully open).

  The first external EGR device 39 is provided in the middle of the first EGR passage 40 and a first EGR passage 40 that communicates the exhaust passage 30 downstream of the turbine 36 and the intake passage 20 upstream of the compressor 37. The first EGR valve 41 controls the flow rate of the exhaust gas passing through the interior, and the water-cooled EGR cooler 42 reduces the temperature of the exhaust gas passing through the first EGR passage 40.

  The first EGR passage 40 includes a portion of the common pipe portion 32 of the exhaust passage 30 that is located downstream of the turbine 36 (downstream in the exhaust gas flow direction) and a compressor 37 of the common pipe portion 23 of the intake passage 20. Also communicates with a portion located upstream (upstream in the intake flow direction). For this reason, when the first EGR valve 41 is opened, the exhaust gas after passing through the turbine 36 flows into the first EGR passage 40, and through the passage 40, the intake passage 20 (common pipe portion) upstream of the compressor 37. 23). In other words, the first external EGR device 39 recirculates the exhaust gas having a relatively low pressure after passing through the turbine 36 to the intake passage 20.

  The second external EGR device 44 is provided in the middle of the second EGR passage 45 and a second EGR passage 45 that communicates the exhaust passage 30 upstream of the turbine 36 and the intake passage 20 downstream of the compressor 37. And a second EGR valve 46 for controlling the flow rate of the exhaust gas passing through the interior, and a water-cooled EGR cooler 47 for reducing the temperature of the exhaust gas passing through the second EGR passage 45.

  The second EGR passage 45 includes a portion located on the upstream side (upstream side in the exhaust gas flow direction) of the common pipe portion 32 of the exhaust passage 30 and the compressor 37 of the common pipe portion 23 of the intake passage 20. Also communicates with a portion located on the downstream side (downstream side in the intake flow direction). For this reason, when the second EGR valve 46 is opened, the exhaust gas before passing through the turbine 36 flows into the second EGR passage 45, and through the passage 45, the intake passage 20 (common pipe portion) on the downstream side of the compressor 37. 23). That is, the second external EGR device 44 returns the exhaust gas having a relatively high pressure before passing through the turbine 36 to the intake passage 20.

  An air cleaner 26 is provided on the upstream side of the compressor 37 in the common pipe portion 23 of the intake passage 20 to remove foreign matters contained in the intake air, and the common pipe portion 23 on the downstream side of the compressor 37. Is provided with an intercooler 25 for cooling the intake air whose temperature has increased due to supercharging.

  In addition, a throttle valve 24 is provided on the downstream side of the intercooler 25 in the common pipe portion 23 of the intake passage 20 so as to be opened and closed. However, in the engine of the present embodiment, an operation (internal EGR) in which exhaust gas remains in the combustion chamber 5 by opening the exhaust valve 9 twice (opening during the intake stroke), or the first and second external EGR devices 39. , 44 can be used to recirculate the exhaust gas into the intake passage 20 (external EGR), and furthermore, the supercharging amount can be adjusted by the operation of the supercharger 35. Based on this, the amount of air (fresh air) introduced into the combustion chamber 5 can be adjusted without operating the throttle valve 24. For this reason, the throttle valve 24 is basically kept fully open except when the engine is stopped.

  A catalytic converter 28 for purifying exhaust gas is provided downstream of the turbine 36 in the common pipe portion 32 of the exhaust passage 30. A three-way catalyst is built in the catalytic converter 28, and harmful components in the exhaust gas passing through the exhaust passage 30 are purified by the action of the three-way catalyst.

(2) Control System FIG. 3 is a block diagram showing an engine control system. The ECU 50 shown in the figure is a device (control means according to the present invention) for comprehensively controlling each part of the engine, and is composed of a well-known CPU, ROM, RAM, and the like.

  Various information is input to the ECU 50 from various sensors provided in the engine. That is, the ECU 50 is electrically connected to the water temperature sensor SW1, the crank angle sensor SW2, and the cam angle sensor SW3 provided in the engine, and based on the input signals from these sensors SW1 to SW3, Various information such as the coolant temperature, crank angle, engine rotation speed, and cylinder discrimination information is acquired.

  The ECU 50 also receives information from various sensors provided in the vehicle. For example, the vehicle is provided with an accelerator opening sensor SW4 that detects the opening degree (accelerator opening degree) of an accelerator pedal (not shown) that is depressed by the driver, which is detected by the accelerator opening degree sensor SW4. The accelerator opening is input to the ECU 50.

  A more specific function of the ECU 50 will be described. The ECU 50 includes, as main functional elements, a determination unit 51, an injector control unit 52, an intake valve control unit 53, an internal EGR control unit 54, and an external EGR control unit. 55, ignition control means 56, and supercharging control means 57.

  The determination means 51 determines the engine based on the engine speed specified from the detected value of the crank angle sensor SW2 and the engine load (target torque) specified from the detected value of the accelerator opening sensor SW4. It is determined each time whether it should be controlled in a manner. In the following, it is assumed that the engine rotation speed is Ne and the engine load is T.

  4 and 5 show an example of a region determination map for determining the engine control content based on the engine rotation speed Ne and the load T. FIG. Note that the map in FIG. 4 is for a warm state where the coolant temperature detected by the engine coolant temperature sensor SW1 is high (eg, 80 ° C. or higher), and the map in FIG. 5 is a cold state where the coolant temperature of the engine is low. It is for the state.

  First, the warm map (FIG. 4) will be described. In the map of FIG. 4, the first operation region A1 is set over the entire rotational speed region in a region where the engine load T is relatively low (low load region). Further, in the region on the higher load side (mainly the middle load region) than the first operation region A1, the second operation region A2 is set over the entire rotational speed region. Furthermore, a third operation region A3 and a fourth operation region A4 are set in order from the low rotation side in a region (mainly a high load region) on the higher load side than the second operation region A2. That is, in the region on the higher load side than the second operation region A2, the third operation region A3 is set in a region where the rotational speed Ne is lower than a predetermined value (for example, about 2000 to 3000 rpm). The fourth operation region A4 is set in a region where the rotational speed Ne is higher than A2.

  During operation of the engine, which operation region (A1 to A4) in FIG. 4 corresponds to the operation point of the engine (the point on the map specified from each value of the load T and the rotational speed Ne). It is judged each time and appropriate control according to each operation region is executed.

  The contents of the control based on the map of FIG. 4 will be briefly described. In this map, the partial load areas excluding the third and fourth operation areas A3 and A4, that is, the first operation area A1 and the second operation area A2 are both mixed with the air-fuel mixture by the compression action of the piston 4. It is set as an execution region of HCCI combustion (compression self-ignition combustion) for self-ignition. However, in the first operation region A1, the operation by the natural intake (Natural Aspiration) in which the supercharger 35 is substantially stopped is performed, while in the second operation region A2, the supercharger 35 performs the supercharge. . Therefore, hereinafter, the control in the first and second operation areas A1 and A2 will be referred to as “NA-HCCI mode” and “supercharged HCCI mode”, respectively.

  On the other hand, the third and fourth operation regions A3 and A4 set on the higher load side than the first and second operation regions A1 and A2 are not HCCI combustion but spark ignition (Spark Ignition) using the spark plug 11. ) Is set as an execution region of SI combustion in which the air-fuel mixture is combusted by flame propagation. However, SI combustion in the third and fourth operation regions A3 and A4 is different from general SI combustion, and the air-fuel mixture is rapidly propagated while the fuel injection timing and ignition timing are set later. It is to burn (the details will be described later). Further, in the third operation region A3 and the fourth operation region A4, the external EGR that recirculates the exhaust gas from the exhaust passage 30 to the intake passage 20 is executed. For the external EGR, the first external EGR device 39 or Which of the second external EGR devices 44 is used is different. Hereinafter, the control in the third operation region A3 is referred to as “first retard SI mode”, and the control in the fourth operation region A4 is referred to as “second retard SI mode”.

  Next, the cold map shown in FIG. 5 will be described. The fifth operation region B1 is set in the low load region of the engine in this map, and the sixth operation region is set in the middle and high load region where the load T is higher than this. The operation area B2 is set. When the map (FIG. 5) is cold, unlike the above-described warm time (FIG. 4), HCCI combustion is not performed, and SI combustion is performed in any operating region. However, supercharging by the supercharger 35 is stopped in the fifth operation region B1 on the low load side, and supercharging by the supercharger 35 is performed in the sixth operation region B2 on the high load side. Hereinafter, the control in the fifth operation region B1 is referred to as “NA-SI mode”, and the control in the sixth operation region B2 is referred to as “supercharging SI mode”.

  Returning to FIG. 3 again, the injector control means 52 electromagnetically opens and closes an unillustrated needle valve (a valve that opens and closes the nozzle hole at the tip of the injector 10) built in the injector 10. The injection amount and injection timing of the fuel injected from 10 to the combustion chamber 5 are controlled.

  The intake valve control means 53 executes control for changing the lift characteristics (opening / closing timing, etc.) of the intake valve 8 by driving the variable mechanism 17.

  The internal EGR control means 54 drives the switching mechanism 18 to execute or stop the valve opening (twice opening) during the intake stroke of the exhaust valve 9, thereby leaving the exhaust gas in the combustion chamber 5 (back flow). The operation to be performed (internal EGR) is switched.

  The external EGR control means 55 recirculates exhaust gas from the exhaust passage 30 to the intake passage 20 by controlling the opening degree of the EGR valves 41 and 46 of the first external EGR device 39 and the second external EGR device 44. The operation (external EGR) to be performed is switched, and the exhaust gas recirculation amount (external EGR amount) is adjusted when the external EGR is executed.

  The ignition control means 56 controls the timing of spark ignition (ignition timing) by the spark plug 11 and the like. However, in the present embodiment, the spark plug 11 operates only when the engine is operated by SI combustion or when the ignition assist that assists the self-ignition of the air-fuel mixture is necessary, and basically otherwise. It does not operate (except for the ignition operation during the intake stroke and exhaust stroke performed to remove carbon).

  The supercharging control means 57 controls the opening degree of the wastegate valve 49, thereby switching the supercharging device 35 to the presence or absence of supercharging and adjusting the supercharging amount when supercharging is executed.

(3) Combustion mode in each operation region during warming Next, each operation region during warming (A1, A2, A3, A4) shown in FIG. 4 based on the control of the ECU 50 having the above functions. Now, what kind of combustion mode is selected will be described. When the engine is operating in a warm state, the ECU 50 determines that the engine operating point (load T and rotational speed Ne) is as shown in FIG. 4 based on the detected values of the crank angle sensor SW2 and the accelerator opening sensor SW4. It is sequentially determined which operation region corresponds to the control map. And the following control is performed according to which of the determined driving | operation area | region is in any of the 1st-4th driving | running area | region A1-A4 in FIG.

(I) First operation area A1 (NA-HCCI mode)
FIG. 6 is a diagram for explaining the combustion mode when the engine is operated in the NA-HCCI mode because the operating point of the engine is in the first operating region A1. As shown in this figure, in this NA-HCCI mode, general HCCI combustion is performed in which an air-fuel mixture based on fuel and air injected during the intake stroke is self-ignited by the compression action of the piston 4.

  Specifically, in the present embodiment, when operating in the NA-HCCI mode, a relatively small amount of fuel is injected (P1) from the injector 10 into the combustion chamber 5 at a predetermined time during the intake stroke. Further, as shown in FIG. 9, when the wastegate valve 49 is fully opened, the supercharger 35 is substantially stopped, and air is introduced into the combustion chamber 5 from the intake passage 20 by natural intake air. A homogeneous and lean air-fuel mixture formed based on a small amount of fuel injected by the fuel injection P1 and air (new air) introduced from the intake passage 20 into the combustion chamber 5 is compressed by the piston 4. Due to the action, the temperature is increased and the pressure is increased, and self-ignition occurs near the compression top dead center (TDC between the compression stroke and the exhaust stroke). Then, based on such self-ignition, combustion accompanied by heat generation as shown in the waveform Q1 occurs.

  In the NA-HCCI mode, a value obtained by dividing the air-fuel ratio of the air-fuel mixture in the combustion chamber 5 (actual air-fuel ratio obtained by dividing the mass of air contained in the air-fuel mixture by the mass of fuel) by the theoretical air-fuel ratio (14.7). The excess air ratio λ is set to be 2 or more. However, under such a substantially lean and homogeneous air-fuel ratio, misfire may occur unless the in-cylinder temperature is intentionally increased. Therefore, in the NA-HCCI mode, the switching mechanism 18 is driven to open the exhaust valve 9 during the intake stroke, that is, the exhaust valve 9 is opened twice, so that the exhaust gas generated in the combustion chamber 5 is opened. The internal EGR is caused to flow back to the combustion chamber 5. Thus, by causing the high-temperature exhaust gas to flow backward (remain) in the combustion chamber 5, the combustion chamber 5 is heated to promote self-ignition of the air-fuel mixture.

  The amount of exhaust gas remaining in the combustion chamber 5 due to the internal EGR (internal EGR amount) is set to be larger on the low load side where the ignitability is poor, and conversely is set to be smaller on the high load side. When the variable mechanism 17 has a function of changing the lift amount of the intake valve 8, the lift amount of the intake valve 8 is gradually increased as the load T increases in order to control the internal EGR amount as described above. The increasing operation may be performed using the variable mechanism 17. Alternatively, in this embodiment, two exhaust valves 9 are provided per cylinder, so that the number of exhaust valves 9 that are opened (opened twice) during the intake stroke is switched between 0, 1, and 2. Thus, the internal EGR amount can be changed stepwise. In this case, the internal EGR amount can be adjusted without changing the lift amount of the intake valve 8.

  Of the first operation region A1 in which the NA-HCCI mode is executed, the operation region in which the load T is significantly reduced (close to no load) in which the fuel injection amount is significantly reduced, or the heat reception of the injected fuel. In the high speed region where the period (actual time during which the fuel is exposed to high temperature and high pressure) is shortened, even if the temperature of the combustion chamber 5 is increased by internal EGR, the air-fuel mixture does not ignite stably. There is a fear. Therefore, in such an operation region, when a part of the air-fuel mixture is combusted by auxiliary spark ignition of the spark plug 11 and the remaining air-fuel mixture is combusted by self-ignition triggered by the combustion, so-called ignition assist is performed. Good.

  In the NA-HCCI mode, the internal EGR based on the opening of the exhaust valve 9 as described above is executed, so the external EGR is stopped. That is, the recirculation of the exhaust gas from the exhaust passage 30 to the intake passage 20 is stopped by setting the opening degree of each EGR valve 41, 46 of the first and second external EGR devices 39, 44 to be fully closed. The

(Ii) Second operation area A2 (supercharged HCCI mode)
In the second operation region A2 set in a region where the load T is higher than the first operation region A1, the control as shown in FIG. 7 is executed as the supercharging HCCI mode. That is, in the supercharging HCCI mode, at least one fuel injection P2 is executed at a predetermined time including the intake stroke in order to combust the air-fuel mixture by self-ignition (HCCI combustion) as in the NA-HCCI mode. . Further, as shown in FIG. 10, the wastegate valve 49 is fully closed so that all the exhaust gas flowing through the exhaust passage 30 flows into the turbine 36 of the supercharger 35, and accordingly, the turbine 36 and the compressor 37. Is driven to rotate, so that the intake air is fully charged.

  The excess air ratio λ is 2 or more (λ ≧ 2) based on the air (fresh air) supplied in large quantities by supercharging of the supercharger 35 and the fuel injected by the fuel injection P2. A lean air-fuel mixture is formed, and the air-fuel mixture self-ignites near the compression top dead center due to the compression action of the piston 4, and combustion accompanied by heat generation as shown by a waveform Q2 in FIG. 7 occurs. The second operation region A2 that is the execution region of the supercharged HCCI mode has a higher load T and a larger amount of injected fuel than the first operation region A1 that is the execution region of the NA-HCCI mode. Therefore, the heat generation amount (waveform Q2) here is larger than the heat generation amount (waveform Q1 in FIG. 6) in the first operation region A1.

  As described above, in the second operation region A2 in which a relatively large amount of fuel corresponding to the load is injected, when the fuel is injected once during the intake stroke, the combustion pressure is particularly high in a part of the high load side. There is a risk that a sudden combustion noise will be generated, or that the combustion temperature will be excessively high and the amount of NOx produced will increase. Therefore, in a part on the high load side in the second operation region A2, in addition to the fuel injection P2 during the intake stroke, fuel injection (injection P2 ′ indicated by a broken line in FIG. 7) is executed after the middle of the compression stroke. Good. As described above, when the required amount of fuel is injected in two steps, that is, the intake stroke and the compression stroke, the heat receiving period of the fuel is shortened on average, and in particular, it is delayed after the middle stage of the compression stroke. The temperature of the combustion chamber 5 near the compression top dead center is lowered by the latent heat of vaporization of the second fuel injection P2 ′ executed at the timing of the combustion, so that the combustion is slowed down and the increase in combustion noise and NOx as described above is prevented. Is done.

  In the supercharging HCCI mode, the switching mechanism 18 is driven so as to invalidate the function of depressing the exhaust valve 9 during the intake stroke, and the valve opening (opening twice) during the intake stroke of the exhaust valve 9 is stopped. The As a result, the exhaust gas hardly flows back into the combustion chamber 5, and internal EGR is prohibited.

  On the other hand, in the supercharged HCCI mode, external EGR is executed instead of the prohibited internal EGR as described above. The external EGR here is an external EGR using the first external EGR device 39 in the embodiment at that time. That is, when the EGR valve 41 of the first external EGR device 39 is opened, the exhaust gas having a relatively low pressure after passing through the turbine 36 of the supercharger 35 passes through the first EGR passage 40 to the intake passage 20. Refluxed.

  As described in (i) and (ii) above, in the engine partial load region (first and second operation regions A1 and A2), the air-fuel mixture has an excess air ratio λ of 2 or more in a significantly lean environment. HCCI combustion is performed at (NA-HCCI mode and supercharged HCCI mode). When the air-fuel mixture that is set to be lean in this way is subjected to HCCI combustion, the combustion temperature is greatly lowered, so that the cooling loss can be reduced and the thermal efficiency (fuel consumption) can be improved. When leaning to λ ≧ 2, the NOx purification action by the three-way catalyst can hardly be expected, but the NOx amount (raw NOx amount) generated by low temperature combustion at λ ≧ 2 is greatly reduced. Even if a special catalyst (for example, NOx trap catalyst) is not provided other than the three-way catalyst, the amount of NOx contained in the exhaust gas can be suppressed to a sufficiently small value.

  Here, it is also conceivable that HCCI combustion with λ ≧ 2 as described above is continuously performed in all engine load ranges (up to the maximum engine load). However, as described above, if the engine load T increases to some extent (for example, increases to the high load side of the second operation region A2), combustion is caused by split fuel injection due to problems of combustion noise and NOx generation amount. Need to slow down. For this reason, in order to continue the HCCI combustion at λ ≧ 2 to a higher load side than the second operation region A2, it is necessary to further promote the slowing of the combustion as described above. It is necessary to inject most of the fuel after the middle stage of the compression stroke, and to extremely reduce the effective compression ratio of the engine based on the closing timing of the intake valve 8. Moreover, since the amount of injected fuel is considerably increased in the vicinity of the maximum load of the engine, in order to maintain a lean air-fuel ratio of λ ≧ 2 for such a large amount of fuel, an overload by the supercharger 35 is required. It is necessary to increase the supply amount extremely, and the required performance of the supercharger 35 becomes excessive. If the supercharging amount increases too much, the exhaust pressure of the engine (the pressure of the exhaust gas upstream from the turbine 36) increases excessively, and as a result of an increase in pump loss, there is a concern that the thermal efficiency deteriorates.

  Therefore, the present inventor is better to switch to SI combustion at λ = 1 (theoretical air-fuel ratio) instead of HCCI combustion at λ ≧ 2 as described above, mainly in the high load region of the engine. The conclusion was reached. However, simply switching to the conventional SI combustion deteriorates the thermal efficiency in the high load range and diminishes effects such as fuel efficiency improvement. Therefore, the following special SI combustion is performed in the third and fourth operation regions A3 and A4 on the higher load side than the second operation region A2.

(Iii) Third and fourth operation areas A3 and A4 (first and second retarded SI modes)
In the third and fourth operation areas A3 and A4 set in a relatively high load area including the maximum engine load, the retard SI mode (first and second retard SI modes) is as shown in FIG. Control is executed. That is, fuel is injected from the injector 10 at the latter stage of the compression stroke (P3), and spark ignition is performed by the spark plug 11 after the fuel injection P3, so that the timing when the compression top dead center is passed (the initial stage of the expansion stroke) From this, control is performed to burn the air-fuel mixture by flame propagation. In this specification, the term “late stage” or “initial stage” of a certain process refers to the latter period or the initial stage when the process is divided into three stages: initial, middle, and late. For example, in the later stage of the compression stroke, it refers to the range of 60 to 0 ° CA before compression top dead center (BTDC), and in the early stage of the expansion stroke, it is from 0 to 60 ° CA after compression top dead center (ATDC). Will point to the range.

  Further, in the first and second retarded SI modes, external EGR for returning the exhaust gas discharged to the exhaust passage 30 to the intake passage 20 is executed. However, in the first retarded SI mode selected in the third operating region A3 and the second retarded SI mode selected in the fourth operating region A4, the first external EGR is used as a device for performing external EGR. Which of the device 39 and the second external EGR device 44 is used is different.

  Specifically, in the present embodiment, in the third operating region A3 (first retarded SI mode) in which the engine rotational speed Ne is relatively low, as shown in FIG. 10, an external device using the first external EGR device 39 is used. When EGR is executed and in the fourth operating region A4 (second retarded SI mode) on the higher rotation side than the third operating region A3, as shown in FIG. 11, the external EGR using the second external EGR device 44 is Executed. As described above, the selective use of the first and second external EGR devices 39 and 44 depending on the rotational speed Ne is derived from the difference in the flow rate of the exhaust gas passing through the exhaust passage 30.

  That is, in the third operating region A3 on the low rotation side, since the flow rate of the exhaust gas passing through the exhaust passage 30 is small, it is assumed that external EGR using the second external EGR device 44 is performed (see FIG. 11). If the exhaust gas is diverted from the upstream side of the turbine 36 of the supercharger 35, the flow rate of the exhaust gas flowing into the turbine 36 may decrease, and the supercharging capability of the supercharger 35 may not be sufficiently secured. is there. On the other hand, in the fourth operating region A4 on the high rotation side, since the flow rate of the exhaust gas passing through the exhaust passage 30 is large, the exhaust gas is temporarily divided from the downstream side of the turbine 36 using the first external EGR device 39. Assuming that (see FIG. 10), the flow rate of the exhaust gas flowing into the turbine 36 becomes excessive, which may cause an excessive increase in the exhaust pressure of the engine and an increase in pump loss based thereon.

  From the above circumstances, in the present embodiment, the first external EGR device 39 is used in the third operation region A3 (first retard SI mode) (FIG. 10), and the fourth operation region A4 (second retard SI mode). ), The second external EGR device 44 is used (FIG. 11). In the fourth operation region A4 (second retarded SI mode), as shown in FIG. 11, the waste gate is used so that the exhaust pressure of the engine (the pressure of the exhaust gas upstream of the turbine 36) does not increase excessively. The valve 49 is opened within a range of less than 100% (fully open), and a part of the exhaust gas is introduced into the bypass pipe 48 (bypassing the turbine 36).

  The timing of the fuel injection P3 in the first and second retarded SI modes is set to an appropriate timing in the later stage of the compression stroke (BTDC 60 to 0 ° CA). The spark ignition timing in the same mode is set to a relatively early timing in the initial stage of the expansion stroke (ATDC 0 to 60 ° CA), specifically within a range of about ATDC 0 to 20 ° CA. However, in the second retarded SI mode selected in the fourth operating region A4 where the engine speed Ne is relatively high, compared to the first retarded SI mode selected in the third operating region A3 on the low rotation side. The timing of fuel injection P3 and spark ignition is set to the more advanced side in the above crank angle range. This is because the combustion of the air-fuel mixture is started at the same timing as that at the time of low rotation even at high rotation when the amount of change in the crank angle per unit time is large.

  The injection amount by the fuel injection P3 is set to a value such that the average air-fuel ratio of the entire combustion chamber 5 becomes the stoichiometric air-fuel ratio (excess air ratio λ = 1). The stoichiometric air-fuel ratio (λ = 1) air-fuel mixture formed based on the fuel injection P3 has a timing after a relatively short period from the completion of the fuel injection P3 (in the example of FIG. 8, a little of the compression top dead center). After the spark ignition executed in (after), combustion starts by flame propagation that is faster than usual, and combustion is completed by a time not so late in the expansion stroke as shown by a waveform Q3 in FIG. The reason why rapid flame propagation combustion is realized will be described later).

  Next, it will be described more specifically what kind of SI combustion is realized based on the first and second retarded SI modes. Hereinafter, SI combustion realized by the first and second retarded SI modes is referred to as “rapid retarded SI combustion”.

  FIG. 12 shows the heat generation rate (upper) in the case of rapid retarded SI combustion (solid line) in the first and second retarded SI modes and in the case of conventional SI combustion (broken line) in which fuel injection is performed during the intake stroke. FIG. 6 is an explanatory diagram conceptually showing how the reaction progress of the unburned air-fuel mixture differs (below). As a precondition for this comparison, the geometric compression ratio of the engine is both 18. Further, it is assumed that the load T and the rotational speed Ne are the same, and therefore the fuel injection amount is also the same. However, it is assumed that the pressure of fuel injection is significantly higher in the rapid retarded SI combustion than in the conventional SI combustion (for example, the former injection pressure is 40 MPa and the latter injection pressure is 7 MPa).

  First, in the conventional SI combustion, the fuel injection P ′ is executed during the intake stroke. In the combustion chamber 5, a sufficiently homogeneous air-fuel mixture is formed after the fuel injection P 'and before the piston 4 reaches compression top dead center. In this example, spark ignition is executed at a later timing after the compression top dead center, and after that (after a predetermined ignition delay time), combustion by flame propagation is started at time θig ′. . Thereafter, as shown by the broken line waveform in the upper diagram of FIG. 12, the peak of the heat generation rate is reached when a predetermined period has elapsed from the combustion start timing θig ′, and combustion is performed at the time θend ′ when further time has elapsed from there. Is completed.

  Here, the period from the start of fuel injection to the end of combustion can be referred to as a period during which an unburned mixture can exist (an unburned mixture period). As shown by the broken line in the lower diagram of FIG. 12, the reaction of the unburned mixture gradually proceeds during the existence period of the unburned mixture. In conventional SI combustion, the unburned mixture exists for a very long period, and during that time, the reaction of the unburned mixture continues, so the unburned mixture self-ignites during the flame propagation after spark ignition. There was a problem that abnormal combustion, that is, knocking occurred. In particular, on the low speed side of the engine where the real time corresponding to the same crank angle change amount is relatively long, the reaction of the unburned mixture proceeds more and more while the piston 4 compresses the mixture. The reactivity of the unburned mixture exceeds the ignition threshold at a timing earlier than the combustion start timing θig ′ based on the spark ignition (that is, the unburned mixture self-ignites regardless of the spark ignition), This results in pre-ignition.

  From the above, in the high compression ratio engine as in the present embodiment, when conventional SI combustion is applied in a high load region such as the third and fourth operation regions A3 and A4 (that is, during the intake stroke) If the fuel is injected at a very early timing), it can be seen that abnormal combustion such as pre-ignition or knocking cannot be avoided even if the spark ignition timing is adjusted.

  On the other hand, in the rapid retarded SI combustion, as described above, the injection pressure is extremely high, such as 20 MPa or more (for example, 30 to 40 MPa), and the retarded phase is greatly retarded (for example, BTDC 20 to 0 ° CA). Fuel is injected during the period (P3 in FIG. 8). Performing such high-pressure and late-time injection (hereinafter referred to as high-pressure retarded injection) shortens the duration of the unburned mixture and avoids abnormal combustion. That is, as shown in FIG. 13, the existence period of the unburned mixture includes a period required for fuel injection from the injector 10 ((A) injection period) and a combustible mixture around the spark plug 11 after the injection is completed. Is the sum of the period until (B) mixture formation period) and the period ((C) combustion period) until combustion started by ignition ends, that is, (A) + (B) + (C). As will be described below, high pressure retarded injection shortens the injection period, the mixture formation period, and the combustion period, respectively, thereby shortening the existence period of the unburned mixture.

  First, the high injection pressure relatively increases the amount of fuel injected from the injector 10 per unit time. For this reason, when the fuel injection amount is constant, as shown in the lower part of FIG. 13, the higher the injection pressure, the shorter the period (injection period) required to inject the injection amount. Therefore, the high pressure retarded injection in which the injection pressure is set to be significantly higher than the conventional one contributes to shortening the injection period (A).

  Further, the high injection pressure is advantageous for atomization of the injected fuel spray, and the penetration force of the fuel spray is increased (the flight speed is increased). For this reason, the higher the injection pressure, the shorter the time required for fuel evaporation (fuel evaporation time), and the shorter the time required for the spray to reach around the spark plug 11 (spray arrival time). Since the mixture formation period (B) is a period obtained by adding the fuel evaporation time and the spray arrival time, as shown in the lower part of FIG. 13, the higher the injection pressure, the mixture formation period (B). Becomes shorter. Therefore, the high-pressure retarded injection with a high injection pressure contributes to shortening the mixture formation period (B).

  Thus, if the fuel injection period (A) and the mixture formation period (B) can be shortened by the high injection pressure, the fuel injection timing, more precisely, the injection start timing is delayed accordingly. It becomes possible. In FIGS. 12 and 8, the timing of the fuel injection P3 is delayed to the latter stage of the compression stroke from such a background. And high-pressure injection of fuel at a late timing of the latter stage of the compression stroke leads to an increase in turbulent energy during the combustion period.

  That is, when the fuel injection timing is delayed to the latter stage of the compression stroke, the turbulent energy increases as the fuel injection pressure increases. On the other hand, even if fuel is injected into the combustion chamber 5 at a high injection pressure, if the timing is too early (for example, during the intake stroke), the time until the ignition timing is long, or the compression stroke Due to the compression from the piston 4 inside, the turbulence in the combustion chamber 5 is attenuated. For this reason, when fuel injection is performed at an early timing such as during the intake stroke, the turbulent energy during the combustion period significantly decreases regardless of the injection pressure level.

  The higher the turbulent energy during the combustion period, the shorter the combustion period. Therefore, when the injection timing is the latter half of the compression stroke, the combustion period (C) becomes shorter as the injection pressure is higher, as shown in the lower part of FIG. That is, the high pressure retarded injection in which fuel is injected at a high pressure in the latter stage of the compression stroke contributes to shortening the combustion period (C).

  In the lower part of FIG. 13, the combustion period in the case where fuel is injected during the intake stroke at a low injection pressure as in the past is indicated by white dots. As is clear from the comparison with the conventional combustion period, according to the high-pressure retarded injection of this embodiment in which fuel is injected in the latter half of the compression stroke at a high injection pressure of 20 MPa or more, it can be seen that the combustion period can be greatly shortened. .

  In addition, an injector having a large number of twelve injection holes, such as the injector 10 of the present embodiment, is more advantageous because the turbulent flow energy is further increased and the combustion period is shortened.

  Here, the high-pressure retarded injection (fuel injection P3 in FIGS. 8 and 12) may inject a required amount of fuel at one time at a predetermined timing in the latter half of the compression stroke, but twice in the latter half of the compression stroke. The fuel may be injected separately. As in the latter case, when the high-pressure retarded injection is divided into two, acceleration of fuel vaporization and improvement of turbulent energy can be achieved at a higher level.

  That is, when the fuel injection P3 is divided into two times, the first fuel injection leads to securing a relatively long mixture formation period, which is advantageous for fuel vaporization atomization. Further, the timing of the second fuel injection can be set to a further delayed timing because the sufficient fuel-air mixture formation period is ensured by the first fuel injection. This is advantageous for improving the turbulent energy in the combustion chamber 5 and contributes to shortening the combustion period. In this case, the ratio of the first fuel injection to the second combustion injection may be set such that the injection amount of the second fuel injection is larger than the injection amount of the first fuel injection. By doing so, the turbulent energy in the combustion chamber 5 is sufficiently increased, which is advantageous for shortening the combustion period and thus avoiding abnormal combustion.

  Furthermore, a small amount of fuel may be injected during the intake stroke as a preliminary injection before the fuel injection P3 executed in the latter half of the compression stroke. In particular, on the high speed side of the fourth operation region A4, after the fuel injection P3 is executed in the latter half of the compression stroke, the fuel vaporization and mixing time cannot be sufficiently secured. It is conceivable to perform preliminary injection at the same time.

  As described above, the high pressure retarded injection shortens the fuel injection period (A), the mixture formation period (B), and the combustion period (C). As a result, as shown in FIG. The period from the injection start timing θinj to the combustion end timing θend (the existence period of the unburned mixture) can be significantly shortened compared to the conventional case in which fuel injection is performed during the intake stroke. And by shortening the said period, even if the compression ratio is high and the load T is high, the air-fuel mixture can be burned out by proper flame propagation without causing abnormal combustion. That is, as shown in the upper part of FIG. 13, in the conventional SI combustion in which the intake stroke injection is performed at a low injection pressure, the reaction progress of the unburned mixture exceeds the ignition threshold as shown by the white circles. In contrast, abnormal combustion occurs, but SI combustion using high-pressure retarded injection, that is, rapid retarded SI combustion, the reaction progress of the unburned mixture is ignited until the end of combustion, as indicated by the black dots. The progress of the reaction can be suppressed so as not to exceed the threshold value, and abnormal combustion can be avoided.

  Moreover, in the rapid retarded SI combustion, the combustion period (C) is greatly shortened, so that the combustion start timing θig based on spark ignition is delayed to some extent from the compression top dead center as shown in the example of FIG. Even if it is set to the initial stage of the expansion stroke, combustion does not continue slowly until the expansion stroke has progressed considerably, and a decrease in thermal efficiency and output torque can be avoided. Of course, if the ignition timing is further advanced than in the example of FIG. 12, the combustion start timing θig approaches the compression top dead center accordingly, so further improvement in thermal efficiency and output torque can be expected. If it is advanced earlier, knocking is more likely to occur. Therefore, the ignition timing is set to the advance side as much as possible under the restriction that knocking does not occur. For this reason, as described above, the ignition timing is set, for example, within a range of about 0 to 20 ° CA after compression top dead center (ATDC). By setting the ignition timing within such a range, the combustion start timing θig is set at an appropriate timing after the compression top dead center to some extent, and knocking is avoided.

(Iv) Specific Example of Control Accompanying Change in Operating State First, an example of control when only the load T changes in the low speed region of the engine will be described. FIGS. 14A to 14F show that the engine operating point is the first operating region A1 in the region determination map shown in FIG. 4 when the load T varies from a low load to a high load in the low speed region of the engine. → 2nd operation area A2 → 3rd operation area A3, and this causes various control parameters when the engine control mode changes from NA-HCCI mode → supercharged HCCI mode → first retarded SI mode It is a figure which shows a change. Among these, FIG. 14A shows a breakdown of the amount of fuel and gas (fresh air and EGR gas) charged into the combustion chamber 5. 14B shows the opening of the throttle valve 24, FIG. 14C shows the presence or absence of the exhaust opening twice (opening during the intake stroke of the exhaust valve 9), FIG. 14D shows the opening of the first EGR valve 41, (E) shows the opening degree of the second EGR valve 46, and (f) shows the opening degree of the waste gate valve 49, respectively.

  As shown in FIG. 14B, the opening degree of the throttle valve 24 is always kept fully open regardless of the engine load T (that is, regardless of which control mode).

  As shown in FIG. 14C, the exhaust valve 9 is opened not only in the exhaust stroke but also in the intake stroke (turned ON twice) in the NA-HCCI mode, and is in the supercharged HCCI mode and the first retarded SI. During the mode, the valve opening during the intake stroke is prohibited (opened twice), and the exhaust valve 9 is opened only during the exhaust stroke. As a result, the internal EGR that causes the exhaust gas to flow backward (remain) in the combustion chamber 5 is executed only in the NA-HCCI mode, and is prohibited in the supercharged HCCI mode and the first retarded SI mode.

  As shown in FIG. 14D, the opening degree of the first EGR valve 41 is set to be fully closed in the NA-HCCI mode (that is, external EGR is prohibited). On the other hand, when the mode is shifted to the supercharged HCCI mode, the first EGR valve 41 is opened to a predetermined intermediate opening, so that the external EGR using the first external EGR device 39, that is, after passing through the turbine 36 is relatively An operation of returning the exhaust gas having a low pressure to the intake passage 20 is started. Further, when the mode is shifted to the first retarded SI mode, the opening degree of the first external EGR valve 41 is further increased (temporarily set to full open), and thereafter gradually decreased with an increase in the load T. As described above, when the shift from the supercharged HCCI mode to the first retarded SI mode is performed, the opening degree of the first external EGR valve 41 is temporarily increased in the supercharged HCCI mode as shown in FIG. Since the excess air ratio λ is 2 or more (that is, more than twice as much fresh air is introduced into the combustion chamber 5 as the fresh air corresponding to λ = 1), the amount of external EGR can be reduced by that amount. This is because in the 1-retard SI mode, the excess air ratio λ is 1, and it is necessary to fill all of the air other than fresh air corresponding to λ = 1 with the exhaust gas from the external EGR.

  As shown in FIG. 14E, the opening degree of the second EGR valve 46 is always kept fully closed regardless of the engine load T (that is, regardless of which control mode). That is, the external EGR using the second external EGR device 44 is not executed in any of the NA-HCCI mode, the supercharged HCCI mode, and the first retarded SI mode.

  As shown in FIG. 14F, the opening degree of the waste gate valve 49 is fully opened in the NA-HCCI mode, and is fully closed in the supercharging HCCI mode and the first retarded SI mode. Thereby, the supercharger 35 is substantially stopped in the NA-HCCI mode, while the supercharger 35 is driven to pressurize the intake air in the supercharged HCCI mode and the first retarded SI mode. Is done. Thus, in the supercharged HCCI mode and the first retarded SI mode, the total gas amount (total amount of fresh air and EGR gas) charged in the combustion chamber 5 is greater than the maximum value during natural intake (max at NA). Will also increase. However, in the low speed region of the engine illustrated in FIG. 14, the flow rate of the exhaust gas is small, and the amount of supercharging by the supercharger 35 reaches a relatively early level. Therefore, if the load T increases to a certain extent, the amount of charged gas increases. There is no further increase.

  As a result of the various parameter controls as described above, an in-cylinder environment as shown in FIG. 14A is realized. That is, in any of the NA-HCCI mode, the supercharged HCCI mode, and the first retarded SI mode, the fuel injection amount is gradually increased with the increase of the load T, and a fresh air corresponding to λ = 1 corresponding to this increases. At least secured. On the other hand, the breakdown of other gases (excess gas) excluding fresh air corresponding to λ = 1 differs in each of the above modes. For example, in the lean HCCI mode, the surplus gas is composed of internal EGR gas and fresh air, and in the supercharged HCCI mode, the surplus gas is composed of external EGR gas and fresh air, and in the rapid retarded SI mode. Sometimes the surplus gas is composed only of external EGR gas. Thereby, the excess air ratio λ is set to 2 or more in the NA-HCCI mode and the supercharged HCCI mode, and is set to 1 in the first retarded SI mode. In FIG. 14A, the lines of “λ = 1” and “λ = 2” represent the amount of fresh air necessary to set the excess air ratio λ to 1 or 2, and “G The line “/ F = 35” indicates that the gas air-fuel ratio (G / F), which is a value obtained by dividing the total gas mass in the combustion chamber 5 (the mass of fresh air and EGR gas) by the mass of the fuel, is 35. Represents the amount of gas required.

  Next, an example of control when only the load T changes in the high speed region of the engine will be described. FIGS. 15A to 15F show that the engine operating point is the first operating region A1 in the region determination map shown in FIG. 4 when the load T fluctuates from a low load to a high load within the high speed region of the engine. → The second operation area A2 → the fourth operation area A4, the various control parameters when the engine control mode is changed from NA-HCCI mode → supercharged HCCI mode → second retarded SI mode It is a figure which shows a change.

  According to FIGS. 15A to 15F, basically the same control as in the case described in FIGS. 14A to 14F is performed in the NA-HCCI mode and in the supercharged HCCI mode. Is executed. On the other hand, in the second retarded SI mode, unlike the first retarded SI mode shown in FIGS. 14A to 14F, the opening degree of the first EGR valve 41 (FIG. 15D) is fully increased. On the other hand, the opening degree of the second EGR valve 46 (FIG. 15E) is set to be fully open. That is, in the second retarded SI mode, unlike in the first retarded SI mode, external EGR using the first external EGR device 39 is prohibited and switched to external EGR using the second external EGR device 44. . As a result, as the external EGR gas, the exhaust gas having a relatively high pressure before passing through the turbine 36 is recirculated.

  In the second retard SI mode, unlike the first retard SI mode, the opening of the waste gate valve 49 is set to an intermediate opening that is lower than fully open (100%). More specifically, at the time of transition from the supercharged HCCI mode to the second retarded SI mode, the opening degree of the wastegate valve 49 is set to an intermediate opening degree from fully closed (0%) to less than fully opened (100%), Thereafter, the load T is gradually decreased as the load T increases. This is because the exhaust gas flow rate is higher and the exhaust gas energy is higher in the high speed region of the engine illustrated in FIG. 15 than in the low speed region of the engine as shown in FIG. That is, in the operation region where the combustion is performed at λ = 1 with high rotation and high load such as the execution region of the second retarded SI mode (fourth operation region A4) (that is, the operation region where the energy of the exhaust gas is increased), Even if the waste gate valve 49 is not fully closed, a sufficient amount of supercharging by the supercharger 35 is ensured. Therefore, in order to bypass a part of the exhaust gas from the turbine 36, the opening of the waste gate valve 49 is set at an intermediate position. The opening is set. Thus, in the fourth operation region A4 where the energy of the exhaust gas is the highest, the amount of supercharging by the supercharger 35 becomes excessive by setting the opening of the wastegate valve 49 to an intermediate opening that is lower than the full opening. Thus, an increase in pump loss is prevented.

  In the high speed region of the engine, the in-cylinder environment as shown in FIG. 15A is realized by the above various parameter control. As shown in this figure, in each mode of the NA-HCCI mode, the supercharged HCCI mode, and the second retarded SI mode, the type of gas charged in the combustion chamber 5 and the value of the excess air ratio λ will be described first. This is the same as in FIG. However, in the case of FIG. 15A, the flow rate of the exhaust gas is large, and the supercharging amount by the supercharger 35 is sufficiently ensured. Therefore, unlike the case of FIG. The total gas amount in the 2 retard SI mode continues to rise above the maximum value during natural inspiration (NA max). For this reason, the gas air fuel ratio (G / F) in the combustion chamber 5 is maintained at 35 or more over the entire load region.

(4) Combustion mode in each operation region when cold The combustion mode in each operation region (B1, B2) when the engine shown in FIG. 5 is cold will be briefly described. When the engine is cold, the ignitability of the air-fuel mixture is poor, and therefore, SI combustion using spark ignition is executed in all operating regions of the engine, not HCCI combustion that self-ignites the air-fuel mixture.

  Specifically, in the fifth operation region B1 set in the low load region when the engine is cold, SI combustion by natural intake with the supercharger 35 substantially stopped is executed as the NA-SI mode. SI combustion in this NA-SI mode is basically the same as conventional SI combustion. In other words, fuel is injected during the intake stroke, and spark ignition is executed slightly before the compression top dead center for the λ = 1 (theoretical air-fuel ratio) mixture based on this, thereby compressing combustion due to flame propagation. Starts near top dead center.

  On the other hand, in the sixth operation region B2 on the higher load side than the fifth operation region B1, SI combustion is performed while operating the supercharger 35 as the supercharging SI mode. In other words, in the supercharging SI mode, the wastegate valve 49 is closed and the exhaust gas flows into the turbine 36, whereby the intake air is pressurized (supercharging) by the supercharger 35 and the intake stroke is performed. The fuel-air mixture based on the fuel injection in the inside starts the flame propagation combustion from the vicinity of the compression top dead center triggered by the spark ignition.

  However, in the high compression ratio engine in which the geometric compression ratio is set to 14 or more as in this embodiment, abnormal combustion such as pre-ignition and knocking occurs particularly in a high load region even when it is cold. There is a fear. Therefore, at least in a part on the high load side in the sixth operation region B2 in which the supercharging SI mode is executed, the variable mechanism 17 is driven to close the intake valve 8 as shown in FIG. The timing (IVC) is changed to a timing earlier than the intake bottom dead center (the bottom dead center BDC between the intake stroke and the compression stroke). As a result, the effective compression ratio of the engine is set to a value smaller than the geometric compression ratio, and the compression end pressure / temperature (in-cylinder pressure and in-cylinder temperature near the compression top dead center) is reduced. Is less likely to self-ignite and pre-ignition and knocking are avoided.

  The operation for reducing the effective compression ratio of the engine can also be performed by changing the closing timing (IVC) of the intake valve 8 to a timing later than the intake bottom dead center (BDC) (for example, X in FIG. 16). . However, in this case, when the intake valve 8 continues to open for a while after the transition to the compression stroke, the air once introduced into the combustion chamber 5 is blown back to the intake port 6, so-called intake blowback. Get up. The blown-back intake air is heated by the heat in the intake port 6 and is again introduced into the combustion chamber 5 in the next intake stroke, so that the temperature of the air in the combustion chamber 5 rises and abnormal combustion is likely to occur. An environment will be created.

  Thus, when the closing timing (IVC) of the intake valve 8 is made later than the bottom dead center (BDC) in order to reduce the effective compression ratio of the engine, compared with the case where it is made earlier than the bottom dead center, The effect of avoiding abnormal combustion is relatively low. Therefore, in order to reliably avoid abnormal combustion, the effective compression ratio can be further reduced by changing the closing timing (IVC) of the intake valve 8 to the more retarded side than the bottom dead center (BDC). Need to do.

  From this point of view, in the present embodiment, the intake valve 8 is closed to reduce the effective compression ratio at least in the high load region (a part on the high load side of the sixth operation region B2) when the engine is cold. The timing is set earlier than the intake bottom dead center. Thereby, compared with the case where the closing timing of the intake valve 8 is made later than the intake bottom dead center, the reduction range of the effective compression ratio can be reduced, and the sacrifice of engine output and fuel consumption can be minimized.

(5) Effects, etc. As described above, in the spark ignition type engine with a supercharger of the present embodiment, the third and fourth operation regions A3, A4 (including at least the high load region when the engine is warm) In FIG. 4), a retard SI mode (first and second retard SI modes) is started in which flame propagation combustion by spark ignition of the spark plug 11 is started after the compression top dead center is passed. In the second operation region A2 where the load T is lower than the execution region (A3, A4), the supercharging HCCI mode is executed in which the supercharger 35 is supercharged and the air-fuel mixture is combusted by self-ignition. In the supercharging HCCI mode, the excess air ratio λ of the air-fuel mixture is set to λ ≧ 2 by introducing a large amount of air into the combustion chamber 5 by supercharging, and the lean air-fuel mixture of λ ≧ 2 is set. In the retarded SI mode, the excess air ratio λ of the air-fuel mixture is set to λ = 1 (theoretical air-fuel ratio) and 20 MPa or more from the injector 10 is controlled. run the fuel injection by the injection pressure in the later stage of the compression stroke, further by performing the spark ignition to Rise Zhang stroke Initial by the spark plug 11, the flame propagation of the air-fuel mixture to the compression top dead center after waiting more than a predetermined time period Control for rapid combustion is performed. According to such a configuration, there is an advantage that combustion with high thermal efficiency and excellent emission characteristics can be continued properly to a high load range while properly using HCCI combustion and SI combustion.

  That is, in the above-described embodiment, an air-fuel mixture of λ ≧ 2 based on a large amount of air supercharged by the supercharger 35 is generated by self-ignition in a predetermined load region (second operation region A2) when the engine is warm. Combustion (HCCI combustion) suppresses the combustion temperature sufficiently low, so that the NOx amount (raw NOx amount) generated by combustion can be greatly reduced and the thermal efficiency can be effectively improved.

  On the other hand, on the higher load side (third and fourth operation regions A3, A4) than the execution region of the HCCI combustion when the engine is warm, high-pressure injection (20 MPa or more) at a later timing after the later stage of the compression stroke ( By performing the fuel injection P3) and performing spark ignition on the λ = 1 mixture based on the fuel injection P3), the spark ignition is triggered and the compression top dead center is passed (that is, the in-cylinder temperature and pressure are reduced). The air-fuel mixture can be rapidly burned by flame propagation (SI combustion) after being reduced to some extent. Therefore, unlike conventional SI combustion in which fuel is injected at an early timing such as the intake stroke, it is possible to realize combustion with excellent thermal efficiency with a short combustion period while reliably avoiding abnormal combustion such as pre-ignition and knocking. Can do. In particular, since the excess air ratio λ of the air-fuel mixture is set to λ = 1 (theoretical air-fuel ratio), unlike the case where a lean air-fuel ratio of λ ≧ 2 is maintained, the supercharging amount of intake air is extremely increased. There is no need to increase the pump loss due to excessive increase in the exhaust pressure of the engine and the deterioration of the fuel consumption based on it. Further, if combustion is performed at λ = 1, NOx can be sufficiently purified only by the three-way catalyst built in the catalytic converter 28, and the emission property is also ensured satisfactorily.

  In the above embodiment, since the injector 10 is provided with a multi-hole type injector having a large number (twelve in this embodiment) of nozzle holes, the injector 10 is in the retarded SI mode. Since the turbulent energy generated based on the fuel injection P3 of the fuel is further increased and the vaporization of the fuel is further promoted, the combustion period can be further shortened and the thermal efficiency can be improved.

  In the above embodiment, in the supercharged HCCI mode execution region (second operation region A2), external EGR using the first external EGR device 39 (operation for recirculating exhaust gas from the exhaust passage 30 to the intake passage 20). Therefore, the advantage is that the amount of NOx generated can be effectively reduced by further reducing the combustion temperature of the air-fuel mixture while appropriately securing the supercharging amount for realizing a lean air-fuel ratio of λ ≧ 2. There is.

  Further, in the above-described embodiment, the air-fuel mixture is not automatically supercharged by the supercharger 35 on the lower load side (first operation region A1) than the supercharging HCCI mode execution region (second operation region A2). The NA-HCCI mode in which combustion is performed by ignition is executed, and in this mode, internal EGR is performed in which high-temperature exhaust gas generated by combustion remains in the combustion chamber 5. According to such a configuration, in the operation region where the engine load T is low and the ignitability of the air-fuel mixture is poor, the air-fuel mixture is surely combusted by self-ignition by making full use of the internal EGR to increase the temperature of the combustion chamber 5. Can be made.

  Further, in the above embodiment, the first external EGR device 39 is in the third operation region A3 set to a part on the low rotation side in the execution region (third and fourth operation regions A3, A4) of the retard SI mode. The external EGR using the second external EGR device 44 is executed in the fourth operation region A4 on the higher rotation side than the third operation region A3. In this way, in the third operation region A3 where the rotational speed Ne is low and the flow rate of the exhaust gas is small, the exhaust gas after passing through the turbine 36 is recirculated to the intake passage 20 using the first external EGR device 39 while rotating. If the exhaust gas before passing through the turbine 36 is recirculated to the intake passage 20 using the second external EGR device 44 in the fourth operation region A4 where the speed Ne is high and the flow rate of the exhaust gas is large, the low There is an advantage that it is possible to achieve both the supercharging performance on the rotation side and the reduction of the pump loss on the high rotation side.

  That is, in the third operating region A3 on the low speed side where the flow rate of the exhaust gas is small, the amount of the exhaust gas flowing into the turbine 36 is reduced by returning the exhaust gas after passing through the turbine 36 to the intake passage 20. This prevents a reduction in supercharging capability. On the other hand, in the fourth operation region A4 on the high rotation side where the flow rate of exhaust gas is large, exhaust gas after passing through the turbine 36 is recirculated to the intake passage 20 so that a lot of exhaust gas flows unnecessarily into the turbine 36. This prevents an increase in pump loss.

  In the above embodiment, when the engine is in a cold state, the air-fuel mixture is sparked while the supercharger 35 is supercharged in the sixth operation region B2 (FIG. 5) excluding the low load region of the engine. The supercharging SI mode in which combustion is performed by ignition is executed, and the variable mechanism 17 is driven and the closing timing of the intake valve 8 is at least a part on the high load side in the supercharging SI mode execution region (B2). The timing is set earlier than the bottom dead center. According to such a configuration, even when the engine is cold, in a situation where abnormal combustion (pre-ignition or knocking) is likely to occur because the load T is high, the closing timing of the intake valve 8 is set to be lower than the intake bottom dead center. As a result, the effective compression ratio of the engine can be reduced to prevent abnormal combustion. Moreover, unlike the case where the closing timing of the intake valve 8 is made later than the intake bottom dead center due to a decrease in the effective compression ratio, the combustion chamber is formed by the air blown back to the intake port 6 being introduced into the combustion chamber 5 again. 5 can be avoided, and the effect of preventing abnormal combustion can be further enhanced. As a result, the reduction range of the effective compression ratio can be kept low, and the sacrifice of engine output and fuel consumption can be minimized.

  Although not particularly mentioned in the above embodiment, a bypass pipe for bypassing the EGR cooler 42 may be provided in the EGR passage 40 of the first external EGR device 39. When exhaust gas is recirculated to the intake passage 20 through this bypass pipe, the exhaust gas is introduced into the combustion chamber 5 while maintaining a relatively high temperature. If such an operation is performed, the ignitability of the air-fuel mixture can be improved. For example, in a part on the high speed side in the execution region (second operation region A2) of the supercharging HCCI mode, the heat receiving period of fuel (actual time during which the fuel is exposed to high temperature and high pressure) is short, and the self-ignition of the mixture does not occur. Relatively difficult to get up. Therefore, when the operation region is shifted to such an operation region, the exhaust gas is recirculated through the bypass pipe as described above without passing through the EGR cooler 42, so that the combustion chamber 5 is heated and the air-fuel mixture is surely self-ignited. be able to.

  Further, in the above embodiment, the engine is warmed in the second operation region A2 set mainly in the middle load region, and on a part of the low rotation side in the region where the load T is higher than the second operation region A2. External EGR using the first external EGR device 39 is executed in the set third operation region A3, the load T is higher than that in the second operation region A2, and the rotational speed Ne is higher than that in the third operation region A3. The external EGR using the second external EGR device 44 is executed in the high fourth operation region A4, but this is only an example, and which external EGR device is used in which operation region, It can be appropriately changed depending on engine characteristics and the like.

  In the above embodiment, the injector 10 is a multi-hole injector, and 12 nozzle holes are provided at the tip thereof. However, the number of the nozzle holes is not limited to 12, but from 12 More or less. However, if the number of injection holes is too small, the concentration of the fuel injected from the injector 10 varies greatly in the circumferential direction. For this reason, it is desirable that the number of nozzle holes be eight or more. If the number of injection holes is eight or more, particularly in the retarded SI mode (first and second retarded SI modes) in which the fuel injection timing is delayed until later in the compression stroke, the fuel injection from the injector 10 is performed. Thereafter, an air-fuel mixture having a substantially uniform air-fuel ratio in the circumferential direction can be formed in a very short time, and subsequent combustion can be performed appropriately.

  Moreover, in the said embodiment, although the injection pressure of the fuel from the injector 10 was made constant at 20 Mpa or more, the injection pressure may not be constant and is variably set according to the operation region. Also good. However, even in this case, the fuel injection pressure is set to 20 MPa or more in at least the execution region of the rapid retarded SI mode (third and fourth operation regions A3, A4).

DESCRIPTION OF SYMBOLS 8 Intake valve 10 Injector 11 Spark plug 17 Variable mechanism 20 Intake passage 30 Exhaust passage 35 Supercharger 36 Turbine 37 Compressor 39 1st external EGR device 44 2nd external EGR device 50 ECU (control means)

Claims (6)

A supercharger that pressurizes intake air; an injector that injects fuel into the combustion chamber; an ignition plug that discharges sparks to an air-fuel mixture formed in the combustion chamber based on the fuel injected from the injector; and the supercharger A spark-ignition engine with a supercharger having a geometric compression ratio set to 14 or more, and a control means for controlling the operation of each part including an injector and a spark plug,
The control means executes a retard SI mode in which flame propagation combustion due to spark ignition of the spark plug is started after the compression top dead center at least in a high load region when the engine is warm. In a predetermined load region where the load is lower than the execution region of (2), the supercharger HCCI mode in which the air-fuel mixture is combusted by self-ignition while supercharging the supercharger is executed,
In the supercharged HCCI mode, the excess air ratio λ of the air-fuel mixture is set to λ ≧ 2 by introducing a large amount of air into the combustion chamber by supercharging, and the lean air-fuel mixture of λ ≧ 2 is compressed top dead. Control to burn by self-ignition from near the point is executed,
In the retard SI mode, the air excess ratio of the mixture gas lambda and sets the lambda = 1, the fuel injection by 20MPa or more injection pressure from the injectors perform the compression stroke late, a further spark ignition by the spark plug by executing the Rise Zhang stroke Initial, supercharged spark-ignition engine, wherein a control for rapidly combusted by flame propagating mixture compression top dead center after waiting more than a predetermined period is performed . The spark ignition engine with a supercharger according to claim 1,
The above-mentioned injector is a multi-hole type injector having a plurality of nozzle holes. A spark ignition engine with a supercharger. The spark ignition engine with a supercharger according to claim 1 or 2,
An external EGR device that recirculates exhaust gas discharged from the combustion chamber to the exhaust passage to the intake passage;
In the supercharged HCCI mode, an exhaust gas recirculation operation is executed by the external EGR device. A spark ignition engine with a supercharger. The spark ignition engine with a supercharger according to any one of claims 1 to 3,
The control means executes the NA-HCCI mode in which the air-fuel mixture is combusted by self-ignition without performing supercharging by the supercharger on a lower load side than the execution region of the supercharging HCCI mode,
In the NA-HCCI mode, a spark ignition engine with a supercharger is executed, wherein internal EGR is performed in which high-temperature exhaust gas generated by combustion remains in the combustion chamber. The spark ignition engine with a supercharger according to any one of claims 1 to 4,
The supercharger has a turbine disposed in the exhaust passage of the engine, and a compressor disposed in the intake passage of the engine and driven by the rotation of the turbine,
Between the intake passage and the exhaust passage, a first external EGR device that recirculates the exhaust gas branched from the exhaust passage downstream of the turbine to the intake passage, and the exhaust branched from the exhaust passage upstream of the turbine. A second external EGR device for recirculating gas to the intake passage,
The exhaust gas recirculation operation by the first external EGR device is executed in a part of the low rotation side in the execution region of the retard SI mode, while the second region is executed in the remaining region in the execution region of the retard SI mode. A spark ignition engine equipped with a supercharger, wherein an exhaust gas recirculation operation is performed by an external EGR device. The spark ignition engine with a supercharger according to any one of claims 1 to 5,
A variable mechanism for variably setting at least the closing timing of the intake valve;
The control means executes a supercharging SI mode in which the air-fuel mixture is combusted by spark ignition while supercharging the supercharger in a middle and high load range when the engine is cold. A spark ignition engine with a supercharger, wherein the variable mechanism is driven and the closing timing of the intake valve is set to a timing earlier than bottom dead center at least at a part of the high load side in the execution region.

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