Patent Publication Number: US-10767593-B2

Title: Control system for compression-ignition engine

Description:
TECHNICAL FIELD 
     The present disclosure relates to a control system for a compression-ignition engine, which executes partial compression-ignition combustion in which a mixture gas within a cylinder is partially combusted by spark-ignition (SI combustion) and then the remaining mixture gas is combusted by self-ignition (CI combustion). 
     BACKGROUND OF THE DISCLOSURE 
     Recently, Homogeneous-Charge Compression Ignition (HCCI) combustion in which a gasoline fuel mixed with air is combusted by self-ignition inside a sufficiently compressed combustion chamber has attracted attention. HCCI combustion is a mode in which the mixture gas combusts at a plurality of positions simultaneously without flame propagation, and thus has a higher combustion speed of the mixture gas than in SI combustion (spark-ignition combustion) which is adopted for general gasoline engines. Therefore, HCCI combustion is said to be significantly advantageous in terms of thermal efficiency. However, in a case of implementing HCCI combustion in an engine of an automobile for which improved thermal efficiency is desired, there are various issues to be solved and an engine which suitably performs HCCI combustion has not been put into practical use. That is, while the engine mounted on the automobile greatly changes in operating state and its environmental condition, HCCI combustion has issues such as a combustion start timing of the mixture gas (a timing at which the mixture gas self-ignites) greatly varies due to external factors (e.g., atmospheric temperature) and control during a transient operation in which an engine load sharply changes being difficult. 
     Therefore, instead of combusting all of the mixture gas by self-ignition, it is proposed to combust a portion of the mixture gas by spark-ignition using a spark plug. That is, after forcibly combusting a portion of the mixture gas through flame propagation caused by spark-ignition (SI combustion), the remaining mixture gas is combusted by self-ignition (CI combustion). Hereinafter, such combustion mode is referred to as “SPCCI (SPark Controlled Compression Ignition) combustion.” 
     For example, JP2009-108778A discloses an engine adopting a similar concept to the SPCCI combustion. This engine causes flame propagation combustion by spark-igniting a stratified mixture gas which is formed around a spark plug by a supplementary fuel injection, and then performs a main fuel injection inside a combustion chamber warmed up by an effect of the flame propagation combustion, so as to combust through self-ignition the fuel injected in the main fuel injection. 
     The CI combustion of the SPCCI combustion occurs when a temperature inside a cylinder (in-cylinder temperature) reaches an ignition temperature of the mixture gas determined by a composition of the mixture gas. Fuel efficiency is maximized by causing the CI combustion when the in-cylinder temperature reaches the ignition temperature near a top dead center of compression stroke. The in-cylinder temperature increases as pressure inside the cylinder (in-cylinder pressure) increases. An increase of the in-cylinder pressure on the compression stroke when the SPCCI combustion is performed is caused by two factors: a pressure increase due to a compressing action of a piston and a pressure increase due to heat from the SI combustion. The pressure increase due to the SI combustion becomes more significant as a combustion speed, i.e., a flame propagation speed, is higher. When an amount of burnt gas remaining inside the cylinder (residual gas amount) is large, the flame propagation speed becomes slower since the burnt gas, which is also inert gas, interrupts the flame propagation. Therefore, the pressure increase due to the SI combustion is smaller as the residual gas amount is larger. Thus, in order to control the pressure increase due to the SI combustion so that the CI combustion occurs near TDC of compression stroke, it is considered to increase the combustion speed of the SI combustion by reducing the residual gas amount as an engine load is lower where a fuel amount is smaller and a heat generation amount inside the cylinder is smaller. 
     On the other hand, when the fuel amount is large and the heat generation amount inside the cylinder is large, that is when the engine load is high, a temperature of exhaust gas, i.e., the temperature of the residual gas, increases. Therefore, on the high engine load side, the residual gas amount at high temperature is increased to raise the temperature inside the cylinder and improve stability of the SI combustion. 
     Such an adjustment of the residual gas amount can be considered to be achieved by advancing an open timing of an intake valve on exhaust stroke as the engine load increases (or retarding the open timing as the engine load decreases). For this reason, the present inventors have intensively studied about the open timing of the intake valve. As a result, it was found that by simply advancing the open timing of the intake valve on the exhaust stroke as the engine load increases, another issue arises that excessive combustion noise is generated when the engine load is high. Therefore, to put the SPCCI combustion into practical use, the open timing of the intake valve needs to be controlled more suitably. 
     SUMMARY OF THE DISCLOSURE 
     The present disclosure is made in view of the above situations and aims to provide a control system for a compression-ignition engine, which reduces combustion noise while achieving suitable flame propagation of SI combustion in SPCCI combustion. 
     The present inventors have intensively studied about a cause of the issue in which excessive combustion noise is generated by advancing the open timing of the intake valve as the engine load increases. As a result, it was found that in a condition where the engine load is relatively high and the in-cylinder temperature easily rises, advancing the open timing of the intake valve increases the residual gas (burnt gas remaining inside the cylinder) amount, which raises the in-cylinder temperature when the compression stroke starts due to the heat of the residual gas, and the in-cylinder temperature near TDC of compression stroke excessively rises. Thus, the pressure increase caused by the heat generation of the CI combustion becomes excessive and results in generating excessive combustion noise. Based on this knowledge, the present inventors have discovered that when the engine load is relatively low, the open timing of the intake valve is advanced as the engine load increases so that the residual gas amount for suppressing the pressure increase due to SI combustion reduces as the engine load is lower and the residual gas amount for increasing the in-cylinder temperature increases as the engine load is higher, and when the engine load is relatively high, the open timing of the intake valve is retarded as the engine load increases so that the residual gas amount for increasing the in-cylinder temperature decreases. Thus, the combustion noise is reduced while achieving suitable flame propagation of the SI combustion in SPCCI combustion. 
     According to one aspect of the present disclosure, a control system for a compression-ignition engine is provided. The engine includes a cylinder, an intake passage, an exhaust passage, an intake port communicating the intake passage to the cylinder, an intake valve configured to open and close the intake port, an exhaust port communicating the exhaust passage to the cylinder, an exhaust valve configured to open and close the exhaust port, an injector configured to inject fuel into the cylinder, and a spark plug configured to ignite a mixture gas containing the fuel injected by the injector and air, the engine executing partial compression-ignition combustion in which the mixture gas is spark-ignited with the spark plug to be partially combusted by spark ignition (SI) combustion and the remaining mixture gas self-ignites to be combusted by compression ignition (CI) combustion. The control system includes an intake variable mechanism configured to change an open timing of the intake valve, and a controller including a processor configured to control parts of the engine, including the intake variable mechanism and the spark plug. While the engine is operating within a given first operating range and a second operating range that is on a higher engine load side of the first operating range, the controller controls the intake variable mechanism to form a gas-fuel ratio (G/F) lean environment in which an air-fuel ratio that is a ratio of air to fuel inside the cylinder is near a stoichiometric air-fuel ratio and burnt gas remains inside the cylinder, and controls the spark plug to spark-ignite the mixture gas so as to combust by the partial compression-ignition combustion at a given timing. While the engine is operating within the first operating range, the controller controls the intake variable mechanism to advance, as the engine load increases at a constant engine speed, the open timing of the intake valve on an advancing side of a top dead center of exhaust stroke, and while the engine is operating within the second operating range, the controller controls the intake variable mechanism to retard, as the engine load increases at a constant engine speed, the open timing of the intake valve on the advancing side of the top dead center of the exhaust stroke. 
     According to this configuration, combustion noise is reduced while achieving suitable flame propagation of the SI combustion in SPCCI combustion (partial compression-ignition combustion). 
     For example, in this configuration, within the first operating range on the low load side, the open timing of the intake valve is controlled so as to be retarded, as the engine load decreases, on the advancing side of the top dead center (TDC) of the exhaust stroke (or, so as to advance, as the engine load increases, on the advancing side of TDC of the exhaust stroke). That is, the intake valve is controlled so that an open period of the intake valve during the exhaust stroke becomes shorter as the engine load decreases. 
     Within a low load range in the first operating range, when a large amount of burnt gas (inert gas) remains inside the cylinder, a flame propagation speed becomes slower since the flame propagation is interrupted by the inert gas, and SI combustion easily becomes unstable. In this regard, according to this configuration, within the low engine load range, the amount of burnt gas discharged from the cylinder to the intake port and flowing back into the cylinder again is reduced, and a situation in which the reaction of air and fuel is interrupted by the inert gas is avoided, which improves the stability of SI combustion, that is, the suitable flame propagation of SI combustion in SPCCI combustion is achieved. 
     On the other hand, as the engine load increases, a heat generation amount increases and a temperature of the burnt gas generated inside the cylinder rises. Here, when the in-cylinder temperature is high, flame propagation occurs more easily. In this regard, according to this configuration, within a high load range in the first operating range, the amount of burnt gas discharged from the cylinder to the intake port and flowing back into the cylinder again increases, the in-cylinder temperature increases, and flame propagation is promoted, which improves the stability of SI combustion. 
     However, even within the second operating range that is on a higher engine load side of the first operating range, that is, the range where the temperature of the burnt gas is higher, when the open timing of the intake valve advances as the engine load increases, the in-cylinder temperature becomes excessively high and combustion noise may increase. If combustion noise increases, for example, the ignition timing needs to be retarded so as to retard the start timing of CI combustion. In this case, CI combustion occurs at a timing when the piston descends significantly on the expansion stroke, which decreases fuel efficiency. 
     In this regard, within the second operating range on the higher engine load side, the open timing of the intake valve is retarded on the advancing side of TDC of the exhaust stroke as the engine load increases, which reduces the amount of burnt gas at the high temperature introduced into the cylinder when the engine load is high. Therefore, within the second operating range, the in-cylinder temperature is prevented from being high and an increase in combustion noise is avoided. 
     The intake variable mechanism may simultaneously change the open and close timings of the intake valve. 
     The first and second operating ranges may be adjacent to each other in an engine load direction bordering on a given first reference load. The open timing of the intake valve within the first and second operating ranges may be set so as to continuously change when the engine load changes across the first reference load. 
     According to this configuration, a situation in which the open and close timings of the intake valve greatly vary when an operation point of the engine shifts between the first operating range and the second operating range is prevented. The open timing of the intake valve can reliably be controlled to suitable timings. 
     The control system may further include a booster configured to boost intake air to be introduced into the cylinder. Within the first operating range, the controller may control the booster to not perform the boost. 
     The control system may further include an exhaust gas recirculation (EGR) device including an EGR passage communicating the intake passage to the exhaust passage, and an EGR valve configured to adjust an amount of exhaust gas recirculated into the cylinder from the exhaust passage through the EGR passage. Within at least a portion of an engine speed segment of the second operating range, the controller may control the EGR device to increase an external EGR ratio as the engine load increases, the external EGR ratio being a ratio of the exhaust gas introduced into the cylinder by the EGR device. 
     According to this configuration, within at least the given engine speed segment of the second operating range, the amount of the high-temperature burnt gas (internal EGR gas) remaining inside the cylinder when the engine load is high is reduced, while the amount of the burnt gas (external EGR gas) which is the exhaust gas introduced into the cylinder by the EGR device and a temperature of which is decreased by passing through the EGR passage, is increased. Thus, the temperature inside the cylinder is prevented from being excessively increased while the amount of burnt gas remaining inside the cylinder is secured. 
     The control system may further include an exhaust variable mechanism configured to change a close timing of the exhaust valve. While the engine operating within a third operating range set in a low load segment of an operating range where the partial compression-ignition combustion is performed in the G/F lean environment, the controller may control the exhaust variable mechanism to retard, as the engine load increases, a close timing of the exhaust valve on a retarding side of the top dead center of the exhaust stroke. While the engine operating within a fourth operating range set in a high load segment of the operating range where the partial compression-ignition combustion is performed in the G/F lean environment, the controller may control the exhaust variable mechanism to advance, as the engine load increases, the close timing of the exhaust valve on the retarding side of the top dead center of the exhaust stroke. 
     If the close timing of the exhaust valve is advanced on the retarding side of the top dead center of the exhaust stroke, the amount of burnt gas discharged from the cylinder to the exhaust port and flowing back into the cylinder again reduces. Thus, according to this configuration, within the third operating range set in the low load segment of the operating range where SPCCI combustion is performed in the G/F lean environment and in which the combustion easily becomes unstable due to the low engine load, the amount of burnt gas remaining inside the cylinder when the engine load is low is reduced and the reaction between the fuel and air is promoted to improve combustion stability. Further, within the third operating range, when the engine load is relatively high and the temperature of the burnt gas is relatively high, the amount of this high-temperature burnt gas remaining inside the cylinder (internal EGR gas) is increased to improve combustion stability. 
     Moreover, within the fourth operating range set in the high load segment of the operating range where SPCCI combustion is performed in the G/F lean environment, by advancing the close timing of the exhaust valve on the retarding side of the top dead center of the exhaust stroke as the engine load increases, a situation in which a large amount of burnt gas (internal EGR gas) which becomes high temperature due to the high engine load remains inside the cylinder, on a high engine load side of the fourth operating range, is prevented. Therefore, within this range, the temperature inside the cylinder is prevented from excessively increasing and CI combustion is prevented from starting excessively early, while combustion stability is secured. 
     The first and second operating ranges may be adjacent to each other in the engine load direction bordering on the given first reference load. The third and fourth operating ranges may be adjacent to each other in the engine load direction bordering on a given second reference load. The first and second reference loads may be set to the same value as each other at least in a portion of an engine speed segment. 
     According to this configuration, a situation in which the open and close timings of the intake valve greatly vary when an operation point of the engine shifts between the first operating range and the second operating range, and shifts between the third operating range and the fourth operating range, is prevented. The open timing of the intake valve can be reliably controlled to suitable timings. Further, by adjusting the open timing of the intake valve and the close timing of the exhaust valve, the combustion stability can be more reliably improved within at least a portion of the engine speed segment of the first to fourth operating ranges. 
     During the partial compression-ignition combustion, the controller may set a target SI ratio according to an operating condition of the engine and set an ignition timing of the spark plug based on the target SI ratio, the target SI ratio being a target value of a ratio of a heat amount generated by the SI combustion with respect to a total heat generation amount in one combustion cycle. 
     According to this configuration, by adjusting the ignition timing so as to achieve SPCCI combustion conforming to the target SI ratio, for example, the ratio of CI combustion is increased (i.e., the SI ratio is reduced). This results in improving thermal efficiency by SPCCI combustion as much as possible. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a system diagram schematically illustrating an overall configuration of a compression-ignition engine according to one embodiment of the present disclosure. 
         FIG. 2  shows diagrams illustrating a cross-sectional view of an engine body and a plan view of a piston. 
         FIG. 3  is a schematic plan view illustrating a structure of a cylinder and intake and exhaust systems in the vicinity thereof. 
         FIG. 4  is a block diagram illustrating a control system of the engine. 
         FIGS. 5A to 5C  are operation maps illustrating a difference in control according to a progression of a warm-up of the engine and an engine speed and an engine load, in which  FIG. 5A  is a first operation map used in a warmed-up state,  FIG. 5B  is a second operation map used in a partially warmed-up state, and  FIG. 5C  is a third operation map used in a cold state. 
         FIG. 6  is a flowchart illustrating a procedure for selecting a suitable map from the first to third operation maps. 
         FIG. 7  is a chart illustrating a waveform of a heat generation rate in SPCCI combustion. 
         FIG. 8  shows time charts schematically illustrating a combustion control executed in respective operating ranges of the engine. 
         FIG. 9  is a three-dimensional map illustrating one specific example of an open timing of an intake valve set within a first partially warmed-up range. 
         FIG. 10  is a three-dimensional map illustrating one specific example of a close timing of an exhaust valve set within the first partially warmed-up range. 
         FIG. 11  is a three-dimensional map illustrating one specific example of the open timing of the intake valve set within a first warmed-up range. 
         FIG. 12  is a three-dimensional map illustrating one specific example of the close timing of the exhaust valve set within the first warmed-up range. 
         FIG. 13  is an operation map within the first partially warmed-up range divided into a plurality of sections based on the open and close timings of the intake valve. 
         FIG. 14  is a chart illustrating a relationship between the engine load and the open timing of the intake valve set at respective engine speeds within the first partially warmed-up range. 
         FIGS. 15A to 15D  show charts illustrating the relationship between the engine load and the open timing of the intake valve set at respective engine speeds within the first partially warmed-up range, in which  FIG. 15A  is a chart at a first speed,  FIG. 15B  is a chart at a second speed,  FIG. 15C  is a chart at a third speed, and  FIG. 15D  is a chart at a fourth speed. 
         FIG. 16  is a chart illustrating a relationship between the engine load and the close timing of the intake valve set at respective engine speeds within the first partially warmed-up range. 
         FIG. 17  is an operation map of the first partially warmed-up range divided into a plurality of sections based on the open and close timing of the exhaust valve. 
         FIG. 18  is a chart illustrating a relationship between the engine load and the close timing of the exhaust valve set at respective engine speeds within the first partially warmed-up range. 
         FIGS. 19A to 19D  show charts illustrating the relationship between the engine load and the close timing of the exhaust valve set at respective engine speeds within the first partially warmed-up range, in which  FIG. 19A  is a chart at the first speed,  FIG. 19B  is a chart at the second speed,  FIG. 19C  is a chart at the third speed, and  FIG. 19D  is a chart at the fourth speed. 
         FIG. 20  is an operation map within the first partially warmed-up range divided into a plurality of sections based on a valve overlap period. 
         FIG. 21  is a chart illustrating a relationship between the engine load and the valve overlap period set at respective engine speeds within the first partially warmed-up range. 
         FIG. 22  is a chart illustrating a relationship between the engine load and an external EGR ratio set at respective engine speeds within the first partially warmed-up range. 
         FIGS. 23A to 23D  show charts illustrating the relationship between the engine load and the external EGR ratio set at respective engine speeds within the first partially warmed-up range, in which  FIG. 23A  is a chart at the first speed,  FIG. 23B  is a chart at the second speed,  FIG. 23C  is a chart at the third speed, and  FIG. 23D  is a chart at the fourth speed. 
         FIG. 24  is a chart corresponding to  FIG. 7 , illustrating various defining methods of an SI ratio. 
         FIG. 25  shows charts illustrating changes of the valve overlap period according to an increase of the engine load. 
     
    
    
     DETAILED DESCRIPTION OF THE DISCLOSURE 
     (1) Overall Configuration of Engine 
       FIGS. 1 and 2  are diagrams illustrating a suitable embodiment of a compression-ignition engine (hereinafter, simply referred to as “the engine”) to which a control system of the present disclosure is applied. The engine illustrated in  FIGS. 1 and 2  is a four-cycle gasoline direct-injection engine mounted on a vehicle as a drive source for traveling, and includes an engine body  1 , an intake passage  30  through which intake air to be introduced into the engine body  1  flows, an exhaust passage  40  through which exhaust gas discharged from the engine body  1  flows, and an external exhaust gas recirculation (EGR) device  50  which recirculates a portion of the exhaust gas flowing through the exhaust passage  40  to the intake passage  30 . This external EGR device  50  is one example of an “EGR device.” 
     The engine body  1  has a cylinder block  3  formed therein with cylinders  2 , a cylinder head  4  attached to an upper surface of the cylinder block  3  so as to cover above the cylinders  2 , and a piston  5  reciprocatably fitted into each cylinder  2 . Typically, the engine body  1  is of a multi-cylinder type having a plurality of cylinders (e.g., four cylinders). Here, the description is only given regarding one cylinder  2  for the sake of simplicity. 
     A combustion chamber  6  is defined above the piston  5 , and a fuel containing gasoline as a main component is injected into the combustion chamber  6  by an injector  15  (described later). Further, the supplied fuel is combusted while being mixed with air in the combustion chamber  6 , and expansion force caused by this combustion pushes down the piston  5  and thus it reciprocates in up-and-down directions of the cylinder. Note that the fuel injected into the combustion chamber  6  may be any fuel as long as it contains gasoline as a main component and, for example, it may contain a subcomponent, such as bioethanol, in addition to gasoline. 
     A crankshaft  7 , which is an output shaft of the engine body  1 , is provided below the piston  5 . The crankshaft  7  is connected to the piston  5  via a connecting rod  8  and rotates about its center axis according to the reciprocation (up-and-down motion) of the piston  5 . 
     A geometric compression ratio of the cylinder  2 , that is, a ratio of the volume of the combustion chamber  6  when the piston  5  is at a top dead center (TDC) to the volume of the combustion chamber  6  when the piston  5  is at a bottom dead center (BDC), is set between 13:1 and 30:1, more preferably between 14:1 and 18:1 as a suitable value for SPCCI (SPark Controlled Compression Ignition) combustion described later. More specifically, the geometric compression ratio of the cylinder  2  is set between 14:1 and 17:1 in regular specifications using gasoline fuel having an octane number of about 91, and between 15:1 and 18:1 in high-octane specifications using gasoline fuel having an octane number of about 96. 
     The cylinder block  3  is provided with a crank angle sensor SN 1  which detects a rotational angle of the crankshaft  7  (crank angle) and a rotational speed of the crankshaft  7  (engine speed), and a water temperature sensor SN 2  which detects a temperature of a coolant flowing through inside the cylinder block  3  and the cylinder head  4  (engine water temperature). 
     The cylinder head  4  is formed with an intake port  9  which opens into the combustion chamber  6  to communicate with the intake passage  30  and an exhaust port  10  which opens into the combustion chamber  6  to communicate with the exhaust passage  40 , and is provided with an intake valve  11  which opens and closes the intake port  9  and an exhaust valve  12  which opens and closes the exhaust port  10 . Note that as illustrated in  FIG. 2 , the type of valve of the engine of this embodiment is a four-valve type including two intake valves and two exhaust valves. That is, the intake port  9  includes a first intake port  9 A and a second intake port  9 B, and the exhaust port  10  includes a first exhaust port  10 A and a second exhaust port  10 B (see  FIG. 3 ). One intake valve  11  is provided for each of the first and second intake ports  9 A and  9 B, and one exhaust valve  12  is provided for each of the first and second exhaust ports  10 A and  10 B. 
     As illustrated in  FIG. 3 , a swirl valve  18  openable and closable of the second intake port  9 B is provided therein. The swirl valve  18  is only provided in the second intake port  9 B, and not provided in the first intake port  9 A. When such a swirl valve  18  is driven in the closing direction, since a rate of intake air flowing into the combustion chamber  6  from the first intake port  9 A in which the swirl valve  18  is not provided increases, a circling flow circling around an axial line Z of the cylinder (a center axis of the combustion chamber  6 ), i.e., swirl flow is enhanced. Conversely, driving the swirl valve  18  in the opening direction weakens the swirl flow. Note that the intake port  9  of this embodiment is a tumble port formable of a tumble flow. Therefore, the swirl flow formed when closing the swirl valve  18  is an inclined swirl flow mixed with the tumble flow. 
     The intake valve  11  and the exhaust valve  12  are driven to open and close in conjunction with the rotation of the crankshaft  7  by valve operating mechanisms  13  and  14  including a pair of camshafts disposed in the cylinder head  4 . 
     The valve operating mechanism  13  for the intake valve  11  is built therein with an intake variable valve timing mechanism (VVT)  13   a  configured to change open and close timings of the intake valve  11 . Similarly, the valve operating mechanism  14  for the exhaust valve  12  is built therein with an exhaust VVT  14   a  configured to change open and close timings of the exhaust valve  12 . The intake VVT  13   a  (exhaust VVT  14   a ) is a so-called phase-variable mechanism which changes the open and close timings of the intake valve  11  (exhaust valve  12 ) simultaneously and by the same amount. That is, the open and close timings of the intake valve  11  (exhaust valve  12 ) are changed while keeping the open period of the valve. The intake VVT  13   a  is one example of an “intake variable mechanism” and the exhaust VVT  14   a  is one example of an “exhaust variable mechanism.” 
     The open timing of the intake valve  11  is changeable between a given timing on an advancing side of a top dead center (TDC) of exhaust stroke and a given timing on a retarding side of TDC of the exhaust stroke. The open period of the intake valve  11  is set so that when an open timing IVO of the intake valve  11  is at a most advanced timing (a most advanced timing possible), a close timing IVC of the intake valve  11  is set on the retarding side of a bottom dead center (BDC) of intake stroke. Accordingly, the close timing IVC of the intake valve  11  is changed on the retarding side of BDC of the intake stroke. An open timing EVO of the exhaust valve  12  is changeable between a given timing on the advancing side of TDC of the exhaust stroke and a given timing on the retarding side of TDC of the exhaust stroke. 
     Note that the open timing of the intake valve  11  (exhaust valve  12 ) described here is not a timing when its lift increases from zero (0), but a timing when a gas flow between the intake port  9  (exhaust port  10 ) via the intake valve  11  (exhaust valve  12 ) substantially starts to occur. For example, the lift of the intake valve  11  (exhaust valve  12 ) increases at a substantially constant rate from a seated state of the valve (i.e., after passing a ramp part) and then sharply rises. The open timing of the intake valve  11  (exhaust valve  12 ) described here is the timing when the lift sharply rises. Specifically, this timing is when the lift of the intake valve  11  (exhaust valve  12 ) is about 0.14 mm. Similarly, the close timing of the intake valve  11  (exhaust valve  12 ) described here is not a timing when its lift becomes zero (0), but a timing when a gas flow between the intake port  9  (exhaust port  10 ) via the intake valve  11  (exhaust valve  12 ) substantially stops. For example, the lift of the intake valve  11  (exhaust valve  12 ) decreases relatively sharply and then further at a substantially constant rate toward zero (i.e., a so-called ramp part is set). The close timing of the intake valve  11  (exhaust valve  12 ) described here is the timing when the lift sharply drops. Specifically, this timing is when the lift of the intake valve  11  (exhaust valve  12 ) is about 0.14 mm. 
     The cylinder head  4  is provided with the injector  15  which injects the fuel (mainly gasoline) into the combustion chamber  6 , and a spark plug  16  which ignites mixture gas containing the fuel injected into the combustion chamber  6  from the injector  15  and air introduced into the combustion chamber  6 . The cylinder head  4  is further provided with an in-cylinder pressure sensor SN 3  which detects pressure of the combustion chamber  6  (hereinafter, also referred to as “in-cylinder pressure”). 
     As illustrated in  FIG. 2 , on a crown surface of the piston  5 , a cavity  20  is formed by denting a relatively wide area of the piston  5 , including a center part thereof, to the opposite side from the cylinder head  4  (downward). Further, a squish portion  21  comprised of an annular flat surface is formed in the crown surface of the piston  5  radially outward of the cavity  20 . 
     The injector  15  is a multi-port injector having a plurality of nozzle ports at its tip portion, and the fuel is injected radially from the plurality of nozzle ports. “F” in  FIG. 2  indicates fuel spray injected from the respective nozzle ports and, in the example of  FIG. 2 , the injector  15  has ten nozzle ports at an even interval in a circumferential direction thereof. The injector  15  is disposed in a center portion of a ceiling surface of the combustion chamber  6  so that its tip portion opposes to a center portion (a bottom center portion of the cavity  20 ) of the crown surface of the piston  5 . 
     The spark plug  16  is disposed at a somewhat offset position to the intake side with respect to the injector  15 . The tip portion (electrode portion) of the spark plug  16  is located at a position overlapping with the cavity  20  in the plan view. 
     As illustrated in  FIG. 1 , the intake passage  30  is connected to one side surface of the cylinder head  4  to communicate with the intake ports  9 . Air (fresh air) taken in from an upstream end of the intake passage  30  is introduced into the combustion chamber  6  through the intake passage  30  and the intake port  9 . 
     In the intake passage  30 , an air cleaner  31  which removes foreign matters within the intake air, a throttle valve  32  which adjusts a flow rate of intake air, a booster  33  which pumps the intake air while compressing it, an intercooler  35  which cools the intake air compressed by the booster  33 , and a surge tank  36  are provided in this order from the upstream side. 
     An airflow sensor SN 4  which detects the flow rate of intake air, first and second intake air temperature sensors SN 5  and SN 7  which detect the temperature of the intake air, and first and second intake air pressure sensors SN 6  and SN 8  which detect pressure of the intake air are provided in various parts of the intake passage  30 . The airflow sensor SN 4  and the first intake air temperature sensor SN 5  are provided in a portion of the intake passage  30  between the air cleaner  31  and the throttle valve  32 , and detect the flow rate and the temperature of the intake air passing through this portion. The first intake air pressure sensor SN 6  is provided in a portion of the intake passage  30  between the throttle valve  32  and the booster  33  (downstream of a connection port of an EGR passage  51  described later), and detects the pressure of the intake air passing through this portion. The second intake air temperature sensor SN 7  is provided in a portion of the intake passage  30  between the booster  33  and the intercooler  35 , and detects the temperature of intake air passing through this portion. The second intake air pressure sensor SN 8  is provided in the surge tank  36  and detects the pressure of intake air in the surge tank  36 . 
     The booster  33  is a mechanical booster (supercharger) mechanically linked to the engine body  1 . Although the specific type of the booster  33  is not particularly limited, for example, any of known boosters, such as Lysholm type, Roots type, or centrifugal type, may be used as the booster  33 . 
     An electromagnetic clutch  34  electrically switchable of its operation mode between “engaged” and “disengaged” is provided between the booster  33  and the engine body  1 . When the electromagnetic clutch  34  is engaged, a driving force is transmitted from the engine body  1  to the booster  33  to enter a boosting state where boost by the booster  33  is performed. On the other hand, when the electromagnetic clutch  34  is disengaged, the transmission of the driving force is interrupted to enter a non-boosting state where the boost by the booster  33  is stopped. 
     A bypass passage  38  which bypasses the booster  33  is provided in the intake passage  30 . The bypass passage  38  connects the surge tank  36  to the EGR passage  51  described later. A bypass valve  39  is provided in the bypass passage  38 . 
     The exhaust passage  40  is connected to the other side surface of the cylinder head  4  so as to communicate with the exhaust port  10 . Burnt gas (exhaust gas) generated in the combustion chamber  6  is discharged outside through the exhaust port  10  and the exhaust passage  40 . 
     A catalytic converter  41  is provided in the exhaust passage  40 . The catalytic converter  41  is built therein with a three-way catalyst  41   a  which purifies hazardous components contained within the exhaust gas flowing through the exhaust passage  40  (HC, CO, and NO x ), and a GPF (gasoline-particulate filter)  41   b  which captures particulate matter (PM) contained within the exhaust gas. Note that another catalytic converter built therein with a suitable catalyst, such as a three-way catalyst or a NO x  catalyst, may be added downstream of the catalytic converter  41 . 
     A linear O 2  sensor SN 10  which detects the concentration of oxygen contained within the exhaust gas is provided in a portion of the exhaust passage  40  upstream of the catalyst converter  41 . The linear O 2  sensor SN 10  linearly changes its output value according to the oxygen concentration and an air-fuel ratio of the mixture gas is estimatable based on the output value of the linear O 2  sensor SN 10 . 
     The external EGR device  50  has the EGR passage  51  connecting the exhaust passage  40  to the intake passage  30 , and an EGR cooler  52  and an EGR valve  53  provided in the EGR passage  51 . The EGR passage  51  connects a portion of the exhaust passage  40  downstream of the catalytic converter  41  to a portion of the intake passage  30  between the throttle valve  32  and the booster  33 . The EGR cooler  52  cools the exhaust gas recirculated from the exhaust passage  40  to the intake passage  30  through the EGR passage  51  by heat exchange. The EGR valve  53  is provided in the EGR passage  51  downstream of the EGR cooler  52  (the side close to the intake passage  30 ), and adjusts the flow rate of the exhaust gas flowing through the EGR passage  51 . Hereinafter, the exhaust gas recirculated from the exhaust passage  40  into the combustion chamber  6  (cylinder  2 ) through the EGR passage  51  is referred to as the external EGR gas. 
     A pressure difference sensor SN 9  which detects a difference between pressure upstream of the EGR valve  53  and pressure downstream thereof is provided in the EGR passage  51 . 
     (2) Control System 
       FIG. 4  is a block diagram illustrating a control system of the engine. An ECU (electronic control unit)  100  illustrated in  FIG. 4  is a microprocessor which comprehensively controls the engine, and comprised of a well-known processor  101  (e.g., a central processing unit (CPU)) having associated ROM and RAM, etc. 
     The ECU  100  receives detection signals from various sensors. For example, the ECU  100  is electrically connected to the crank angle sensor SN 1 , the water temperature sensor SN 2 , the in-cylinder pressure sensor SN 3 , the airflow sensor SN 4 , the first and second intake air temperature sensors SN 5  and SN 7 , the first and second intake air pressure sensors SN 6  and SN 8 , the pressure difference sensor SN 9 , and the linear O 2  sensor SN 10 , which are described above. The ECU  100  sequentially receives the information detected by these sensors (i.e., the crank angle, the engine speed, the engine water temperature, the in-cylinder pressure, the intake air flow rate, the intake air temperatures, the intake air pressures, the difference in pressure between the upstream and downstream sides of the EGR valve  53 , the oxygen concentration of the exhaust gas, etc.). 
     Further, an accelerator sensor SN 11  which detects an opening of an accelerator pedal controlled by a vehicle driver driving the vehicle is provided in the vehicle, and a detection signal from the accelerator sensor SN 11  is also inputted to the ECU  100 . 
     The ECU  100  controls various components of the engine while executing various determinations and calculations based on the input signals from the various sensors. That is, the ECU  100  is electrically connected to the intake VVT  13   a , the exhaust VVT  14   a , the injector  15 , the spark plug  16 , the swirl valve  18 , the throttle valve  32 , the electromagnetic clutch  34 , the bypass valve  39 , the EGR valve  53 , etc., and outputs control signals to these components based on various calculation results. Note that the ECU  100  as described above is one example of a “controller.” 
     (3) Control According to Operating State 
       FIGS. 5A to 5C  are operation maps illustrating a difference in control according to a progression of a warm-up of the engine and the engine speed and load. In this embodiment, different operation maps Q 1  to Q 3  are prepared corresponding to three stages including a warmed-up state where the warm-up of the engine is completed, a partially warmed-up state where the engine is in process of warming up, and a cold state where the engine is not warmed up. Hereinafter, the operation map Q 1  used in the warmed-up state is referred to as the first operation map, the operation map Q 2  used in the partially warmed-up state is referred to as the second operation map, and the operation map Q 3  used in the cold state is referred to as the third operation map. 
     Note that in the below description, the engine load being high (low) is equivalent to a required torque of the engine being high (low). Further in the below description, phrases like “early stage,” “middle stage,” and “late stage” of a certain stroke or phrases like “early half” and “latter half” of a certain stroke may be used to specify a timing of a fuel injection or a spark-ignition, and these phrases are based on the following definitions. That is, here, three periods formed by evenly dividing any stroke, such as intake stroke or compression stroke, are defined as “early stage,” “middle stage,” and “late stage,” respectively. Therefore, for example, (i) the early stage, (ii) the middle stage, and (iii) the late stage of the compression stroke indicate (i) a range between 180° C.A and 120° C.A before TDC (BTDC) of the compression stroke, (ii) a range between 120° C.A and 60° C.A BTDC, (iii) a range between 60° C.A and 0° C.A BTDC, respectively. Similarly, here, two periods formed by evenly dividing any stroke, such as the intake stroke or the compression stroke, are defined as “early half” and “latter half,” respectively. Therefore, for example, (iv) the early half and (v) the latter half of the intake stroke indicate (iv) a range between 360° C.A and 270° C.A BTDC, and (v) a range between 270° C.A and 180° C.A BTDC, respectively. 
       FIG. 6  is a flowchart illustrating a procedure for selecting a suitable map from the first to third operation maps Q 1  to Q 3 . Once the control illustrated in this flowchart is started, at S 1 , the ECU  100  determines whether (i) the engine water temperature is below 30° C. and (ii) the intake air temperature is below 25° C. are both satisfied, based on the engine water temperature detected by the water temperature sensor SN 2  and the intake air temperature detected by the second intake air temperature sensor SN 7 . 
     If S 1  is YES and it is confirmed that (i) and (ii) are satisfied, i.e., both “engine water temperature&lt;30° C.” and “intake air temperature&lt;25° C.” are satisfied and the engine is in the cold state, the ECU  100  shifts to S 2  to determine the third operation map Q 3  illustrated in  FIG. 5C  as the operation map to be used. 
     On the other hand, if S 1  is NO and it is confirmed that at least one of (i) and (ii) is not satisfied, the ECU  100  shifts to S 3  to determine whether (iii) the engine water temperature is below 80° C. and (iv) the intake air temperature is below 50° C. are both satisfied, based on the engine water temperature detected by the water temperature sensor SN 2  and the intake air temperature detected by the second intake air temperature sensor SN 7 . 
     If S 3  is YES and it is confirmed that (iii) and (iv) are satisfied, i.e., at least one of “engine water temperature≥30° C.” and “intake air temperature≥25° C.” is satisfied, and both “engine water temperature&lt;80° C.” and “intake air temperature&lt;50° C.” are satisfied, which means that the engine is in the partially warmed-up state, the ECU  100  shifts to S 4  to determine the second operation map Q 2  illustrated in  FIG. 5B  as the operation map to be used. 
     On the other hand, if S 3  is NO and it is confirmed that at least one of (iii) and (iv) is not satisfied, i.e., at least one of “engine water temperature≥80° C.” and “intake air temperature≥50° C.” is satisfied, which means that the engine is in the warmed-up state (warm-up completed state), the ECU  100  shifts to S 5  to determine the first operation map Q 1  illustrated in  FIG. 5A  as the operation map to be used. 
     Next, details of controls (a difference in combustion control according to the engine speed/load) defined by the operation maps Q 1  to Q 3  in the cold state, the partially warmed-up state, and the warmed-up state are described, respectively. 
     (3-1) Control in Cold State 
     A combustion control in the cold state of the engine is described with reference to the third operation map Q 3  ( FIG. 5C ). In the cold state of the engine, a control for mixing the fuel with air to form the mixture gas and performing the SI combustion with the mixture gas is executed within an entire operating range C 1 . The explanation of the control in the cold state is omitted since it is similar to the combustion control of a general gasoline engine. 
     (3-2) Control in Partially Warmed-up State 
     A combustion control in the partially warmed-up state of the engine is described based on the second operation map Q 2  ( FIG. 5B ). As illustrated in  FIG. 5B , when the engine is in the partially warmed-up state, the operating range of the engine is mainly divided into three operating ranges B 1  to B 3 . When the three ranges are a first partially warmed-up range B 1 , a second partially warmed-up range B 2 , and a third partially warmed-up range B 3 , the third partially warmed-up range B 3  is a high engine speed range. The first partially warmed-up range B 1  is a low and medium speed, low load range extending on the lower speed side of the third partially warmed-up range B 3 , excluding the high load side. The second partially warmed-up range B 2  is a range other than the first and third partially warmed-up ranges B 1  and B 3  (i.e., a low and medium speed, high load range). 
     (a) First Partially Warmed-up Range 
     Within the first partially warmed-up range B 1 , the SPCCI combustion combining the SI combustion and the CI combustion is performed. The SI combustion is a mode in which the mixture gas is ignited by the spark plug  16  and is then forcibly combusted by flame propagation which spreads the combusting region from the ignition point, and the CI combustion is a mode in which the mixture gas is combusted by self-ignition in an environment increased in temperature and pressure due to the compression of the piston  5 . The SPCCI combustion combining the SI combustion and the CI combustion is a combustion mode in which the SI combustion is performed on a portion of the mixture gas inside the combustion chamber  6  by the spark-ignition performed in an environment immediately before the mixture gas self-ignites, and after the SI combustion, the CI combustion is performed on the remaining mixture gas in the combustion chamber  6  by self-ignition (by the further increase in temperature and pressure accompanying the SI combustion). Note that “SPCCI” is an abbreviation of “SPark Controlled Compression Ignition” and the SPCCI combustion is one example of “partial compression-ignition combustion.” 
     The SPCCI combustion has a characteristic that the heat generation in the CI combustion is faster than that in the SI combustion. For example, as illustrated in  FIG. 7  described later, a waveform of a heat generation rate caused by the SPCCI combustion has a shape in which a rising slope in an early stage of the combustion which corresponds to the SI combustion is shallower than a rising slope caused corresponding to the CI combustion occurring subsequently. In other words, the waveform of the heat generation rate caused by the SPCCI combustion is formed to have a first heat generation rate portion formed by the SI combustion and having a relatively shallow rising slope, and a second heat generation rate portion formed by the CI combustion and having a relatively steep rising slope, which are next to each other in this order. Further, corresponding to the tendency of such a heat generation rate, in the SPCCI combustion, a pressure increase rate (dp/dθ) inside the combustion chamber  6  caused by the SI combustion is lower than that in the CI combustion. 
     When the temperature and pressure inside the combustion chamber  6  rise due to the SI combustion, the unburnt mixture gas self-ignites and the CI combustion starts. As illustrated in  FIG. 7 , the slope of the waveform of the heat generation rate changes from shallow to steep at the timing of self-ignition (that is, the timing when the CI combustion starts). That is, the waveform of the heat generation rate caused by the SPCCI combustion has a flection point at a timing when the CI combustion starts (indicated by an “X 2 ” in  FIG. 7 ). 
     After the CI combustion starts, the SI combustion and the CI combustion are performed in parallel. In the CI combustion, since the combustion speed of the mixture gas is faster than that in the SI combustion, the heat generation rate becomes relatively high. However, since the CI combustion is performed after TDC of compression stroke, the slope of the waveform of the heat generation rate does not become excessive. That is, after TDC of compression stroke, since the motoring pressure decreases due to the piston  5  descending, the rise of the heat generation rate is prevented, which avoids excessive dp/dθ in the CI combustion. In the SPCCI combustion, due to the CI combustion being performed after SI combustion as described above, it is unlikely for the dp/dθ which is an index of combustion noise to become excessive, and combustion noise is reduced compared to performing the CI combustion alone (in the case where the CI combustion is performed on all the fuel). 
     The SPCCI combustion ends as the CI combustion finishes. Since the combustion speed of the CI combustion is faster than that of the SI combustion, the combustion end timing is advanced compared to performing the SI combustion alone (in the case where the SI combustion is performed on all the fuel). In other words, the SPCCI combustion brings the combustion end timing closer to TDC of compression stroke, on the expansion stroke. Thus, the SPCCI combustion improves the fuel efficiency compared to the SI combustion alone. 
     Within the first partially warmed-up range B 1 , when the spark plug  16  performs the ignition (when the mixture gas starts to combust), an environment in which the burnt gas (combusted gas) exists within the combustion chamber  6 , a gas-fuel ratio (G/F) which is a weight ratio between the entire gas (G) and the fuel (F) within the combustion chamber  6  (cylinder  2 ) is increased to be higher than a stoichiometric air-fuel ratio (14.7:1), and an air-fuel ratio (A/F) which is a ratio between the air (A) and the fuel (F) within the combustion chamber  6  (cylinder  2 ) substantially matches the stoichiometric air-fuel ratio (hereinafter, referred to as a G/F lean environment) is formed and a control for performing the SPCCI combustion of the mixture gas is executed. More specifically, the gas-fuel ratio (G/F) is 18:1≤G/F≤50:1. By setting this range, the stability of the SI combustion is secured, the controllability of the start timing of the CI combustion is secured, and combustion noise is also reduced. 
     In order to achieve the SPCCI combustion in such a G/F lean environment, within the first partially warmed-up range B 1 , various components of the engine are controlled by the ECU  100  as follows. 
     The injector  15  performs at least a single fuel injection on the intake stroke. For example, at an operation point P 2  within the first partially warmed-up range B 1 , the injector  15  performs the single fuel injection for supplying the entire amount of fuel to be injected in one cycle, during the intake stroke as illustrated in a chart (b) of  FIG. 8 . 
     The spark plug  16  ignites the mixture gas near TDC of compression stroke. For example, at the operation point P 2 , the spark plug  16  ignites the mixture gas at a slightly advanced timing than TDC of compression stroke. This ignition triggers the SPCCI combustion, a portion of the mixture gas in the combustion chamber  6  is combusted through flame propagation (SI combustion), and then the remaining mixture gas is combusted by self-ignition (CI combustion). 
     The opening of the throttle valve  32  is set so that an air amount equivalent to the stoichiometric air-fuel ratio is introduced into the combustion chamber  6  through the intake passage  30 , i.e., so that the air-fuel ratio (A/F) which is a weight ratio between air (fresh air) and the fuel inside the combustion chamber  6  substantially matches the stoichiometric air-fuel ratio (14.7:1). On the other hand, within the first partially warmed-up range B 1 , the open timing IVO of the intake valve  11 , a close timing EVC of the exhaust valve  12  and the opening of the EGR valve  53  are adjusted so that the external EGR gas and/or the internal EGR gas, which is the burnt gas, flows into (remains inside) the combustion chamber  6 . Thus, within the first partially warmed-up range B 1 , the gas-fuel ratio is increased to be higher than the stoichiometric air-fuel ratio. The internal EGR gas is, within the burnt gas generated inside the combustion chamber  6 , the portion which is not the external EGR gas, in other words, it is not the burnt gas recirculated into the combustion chamber  6  through the EGR passage  51  but gas remaining inside the combustion chamber  6  without being discharged to the EGR passage  51  (including gas returned back to the combustion chamber  6  after being discharged to the intake port  9  and/or the exhaust port  10 ). 
     An opening of the EGR valve  53  is controlled to achieve a target external EGR ratio variably set within a substantial range of 0-40%. Note that the external EGR ratio used here is a weight ratio of exhaust gas recirculated to the combustion chamber  6  through the EGR passage  51  (external EGR gas) to all the gas inside the combustion chamber  6 , and the target external EGR ratio is a target value of the external EGR ratio. The target external EGR ratio within the first partially warmed-up range B 1  will be described later in detail. 
     The intake VVT  13   a  changes the open timing IVO of the intake valve  11  (intake open timing IVO) according to the engine speed and the engine load as illustrated in  FIG. 9 . The exhaust VVT  13   a  changes the close timing EVC of the exhaust valve  12  (exhaust close timing EVC) according to the engine speed and the engine load as illustrated in  FIG. 10 . These  FIGS. 9 and 10  are three-dimensional maps illustrating specific examples of the open timing IVO of the intake valve  11  (the close timing EVC of the exhaust valve  12 ) with respect to the engine speed and the engine load. The open and close timings of the intake valve  11  and the close timing of the exhaust valve  12  within the first partially warmed-up range B 1  will be described later in detail. 
     The booster  33  is in an OFF state when the engine load is below a given boosting load T_t. On the other hand, within the first partially warmed-up range B 1 , the booster  33  is in an ON state when the engine load is above the boosting load T_t. When the booster  33  is in the OFF state, the electromagnetic clutch  34  is disengaged to disconnect the booster  33  from the engine body  1  and fully open the bypass valve  39  so as to stop the boost by the booster  33  (enter a non-boosting state). On the other hand, when the booster  33  is in the ON state, the electromagnetic clutch  34  is engaged to connect the booster  33  to the engine body  1  so as to boost by the booster  33  (enter a boosting state). Here, the opening of the bypass valve  39  is controlled so that the pressure in the surge tank  36  (boosting pressure) detected by the second intake air pressure sensor SN 7  matches a given target pressure determined for each operating condition of the engine (a condition such as the engine speed and the engine load). For example, as the opening of the bypass valve  39  increases, the flow rate of the intake air which flows back to the upstream side of the booster  33  through the bypass passage  38  increases, and as a result, the pressure of the intake air introduced into the surge tank  36 , that is, the boosting pressure, becomes low. By adjusting the backflow amount of the intake air in this manner, the bypass valve  39  controls the boosting pressure to the target pressure. 
     Within the first partially warmed-up range B 1 , the opening of the swirl valve  18  is adjusted to form a relatively weak swirl flow. For example, the swirl valve  18  is set to be about half open (50%) or have a larger opening. 
     (b) Second Partially Warmed-up Range 
     Within the second partially warmed-up range B 2 , the control for performing the SPCCI combustion of the mixture gas is executed in the environment in which the air-fuel ratio inside the combustion chamber  6  is slightly richer (an excess air ratio λ≤1) than the stoichiometric air-fuel ratio. In order to achieve the SPCCI combustion in such a rich environment, within the second partially warmed-up range B 2 , various components of the engine are controlled by the ECU  100  as follows. 
     The injector  15  injects all or a majority of the fuel for one combustion cycle, during the intake stroke. For example, at an operation point P 3  within the second partially warmed-up range B 2 , the injector  15  injects the fuel over a continuous period overlapping with a latter half of the intake stroke, more specifically, a continuous period from the latter half of the intake stroke to an early half of the compression stroke, as illustrated in the chart (c) of  FIG. 8 . 
     The spark plug  16  ignites the mixture gas near TDC of compression stroke. For example, at the operation point P 3 , the spark plug  16  ignites the mixture gas at a slightly retarded timing than TDC of compression stroke. 
     The booster  33  is controlled to be ON and performs the boost. The boosting pressure here is adjusted by the bypass valve  39 . 
     The intake VVT  13   a  and the exhaust VVT  14   a  set valve operation timings of the intake and exhaust valves  11  and  12  so that the internal EGR gas does not remain inside the combustion chamber  6  (the internal EGR is substantially stopped). The throttle valve  32  is fully opened. The opening of the EGR valve  53  is controlled so that the air-fuel ratio (A/F) in the combustion chamber  6  becomes the stoichiometric air-fuel ratio or slightly richer (λ≤1). For example, the EGR valve  53  adjusts the amount of the exhaust gas recirculated through the EGR passage  51  (external EGR gas) so that the air-fuel ratio becomes between 12:1 and 14:1. Note that near the highest engine load, the EGR valve  53  may be closed to substantially stop the external EGR. The swirl valve  18  is set to have an intermediate opening which is larger than that within the first partially warmed-up range B 1  but smaller than a largest (full) opening. 
     (c) Third Partially Warmed-up Range 
     Within the third partially warmed-up range B 3 , a relatively traditional SI combustion is performed. In order to achieve the SI combustion, within the third partially warmed-up range B 3 , various components of the engine are controlled by the ECU  100  as follows. 
     The injector  15  at least injects the fuel over a given period overlapping with the intake stroke. For example, at an operation point P 4  within the third partially warmed-up range B 3 , the injector  15  injects the fuel over a continuous period from the intake stroke to the compression stroke, as illustrated in the chart (d) of  FIG. 8 . 
     The spark plug  16  ignites the mixture gas near TDC of compression stroke. For example, at the operation point P 4 , the spark plug  16  ignites the mixture gas at a slightly advanced timing than TDC of compression stroke. Further, this ignition triggers the SI combustion, and all of the mixture gas in the combustion chamber  6  combusts through flame propagation. 
     The booster  33  is controlled to be ON and performs the boost. The boosting pressure here is adjusted by the bypass valve  39 . The throttle valve  32  is fully opened. The opening of the EGR valve  53  is controlled so that the air-fuel ratio (A/F) in the combustion chamber  6  becomes the stoichiometric air-fuel ratio or slightly richer (λ≤1). The swirl valve  18  is fully opened. Thus, not only the first intake port  9 A is but also the second intake port  9 B is fully opened and charging efficiency of the engine improves. 
     (3-3) Control in Warmed-up State 
     As illustrated in  FIG. 5A , when the engine is in the warmed-up state, the operating range of the engine is mainly divided into four operating ranges A 1  to A 4 . When the four operating ranges are a first warmed-up range A 1 , a second warmed-up range A 2 , a third warmed-up range A 3  and a fourth warmed-up range A 4 , the second warmed-up range A 2  corresponds to a high load segment of the first partially warmed-up range B 1 , the first warmed-up range A 1  corresponds to the first partially warmed-up range B 1  without the second warmed-up range A 2 , the third warmed-up range A 3  corresponds to the second partially warmed-up range B 2 , and the fourth warmed-up range A 4  corresponds to the third partially warmed-up range B 3 . 
     (a) First Warmed-up Range 
     Within the first warmed-up range A 1 , a control is executed in which SPCCI combustion of the mixture gas is performed while setting the A/F higher than the stoichiometric air-fuel ratio (14.7:1), so as to keep an amount of NO x  generated by the combustion small and improve fuel efficiency. That is, SPCCI combustion is performed while setting the excess air ratio λ&gt;1 inside the combustion chamber  6 . The A/F within the first warmed-up range A 1  is set variably, for example within a range of 20 to below 35, so that the amount of NO x  generated by the combustion is kept sufficiently small. A target air-fuel ratio within the first warmed-up range A 1  is generally set to be higher as the engine load (required torque) increases. 
     In order to achieve the SPCCI combustion in such an environment where the air-fuel ratio is higher than the stoichiometric air-fuel ratio (hereinafter, suitably referred to as an “A/F lean environment”), within the first warmed-up range A 1 , various components of the engine are controlled by the ECU  100  as follows. 
     The injector  15  injects the fuel by splitting it into a plurality of injections from the intake stroke to the compression stroke. For example, at an operation point P 1  at which the engine speed and load are relatively low within the first warmed-up range A 1 , the injector  15  injects the majority of the fuel for one cycle in two portions from an early stage to a middle stage of the intake stroke and the remaining fuel in a final stage of the compression stroke (a total of three injections), as illustrated in the chart (a) of  FIG. 8 . 
     The spark plug  16  ignites the mixture gas near TDC of compression stroke. For example, at the operation point P 1 , the spark plug  16  ignites the mixture gas at a slightly advanced timing than TDC of compression stroke. This ignition triggers SPCCI combustion, a portion of the mixture gas in the combustion chamber  6  is combusted through flame propagation (SI combustion), and then the remaining mixture gas is combusted by self-ignition (CI combustion). 
     The booster  33  is in the OFF state within the substantially entire first warmed-up range A 1 . The throttle valve  32  is fully opened or has a similar opening within the entire first warmed-up range A 1 . Thus, a large amount of air is introduced into the combustion chamber  6  to increase the air-fuel ratio inside the combustion chamber  6 . 
     The intake VVT  13   a  changes the open timing IVO of the intake valve  11  according to the engine speed and the engine load as illustrated in  FIG. 11 . 
     Specifically, substantially within a low load range where the engine load is low, the open timing IVO of the intake valve  11  is advanced as the engine load increases. For example, the intake open timing IVO is set to be retarded than TDC of the exhaust stroke at a lowest engine load and is advanced to a most advanced timing as the engine load increases. Further, within a medium load range where the engine load is relatively high, the intake open timing IVO is kept at the most advanced timing regardless of the engine load. Moreover, within a high load range where the engine load is even higher, the intake open timing IVO is retarded as the engine load increases on a more advancing side of TDC of the exhaust stroke. Note that similar to the intake open timing IVO, the close timing IVC of the intake valve  11  is changed with respect to the engine load on a more retarding side of the BDC of the intake stroke. 
     The exhaust VVT  14   a  changes the close timing EVC of the exhaust valve  12  according to the engine speed and the engine load as illustrated in  FIG. 12 . 
     Specifically, the exhaust close timing EVC is set on the retarding side of TDC of the exhaust stroke. Further, within the low load range, the exhaust close timing EVC is retarded as the engine load increases. For example, the exhaust close timing EVC is set to TDC of the exhaust stroke at the lowest engine load, and its retarded amount from TDC of the exhaust stroke is increased as the engine load increases. Moreover, within the medium load range, the exhaust close timing EVC is kept fixed regardless of the engine load. Furthermore, within the high load range, the exhaust close timing EVC is advanced as the engine load increases. Note that the open timing EVO of the exhaust valve  12  is changed with respect to the engine load similarly to the exhaust close timing EVC. 
     The opening of the EGR valve  53  is controlled to achieve a target external EGR ratio variably set within a substantial range of 0-20%. The target external EGR ratio is increased as the engine speed or the engine load increases. 
     Within the first warmed-up range A 1 , the opening of the swirl valve  18  is set smaller than the half-opened state (50%). By reducing the opening of the swirl valve  18  as above, majority of the intake air introduced into the combustion chamber  6  is from the first intake port  9 A (the intake port on the side where the swirl valve  18  is not provided), and a strong swirl flow is formed inside the combustion chamber  6 . This swirl flow grows during the intake stroke and remains until the middle of the compression stroke, to promote stratification of the fuel. That is, a concentration difference that the fuel in the center portion of the combustion chamber  6  concentrates more than outside thereof (outer circumferential portion) is formed. For example, within the first warmed-up range A 1 , the air-fuel ratio in the center portion of the combustion chamber  6  is set between 20:1 and 30:1 by the effect of the swirl flow, and the air-fuel ratio in the outer circumferential portion of the combustion chamber  6  is set to 35:1 or higher. Within the first warmed-up range A 1 , a target swirl opening is variably set to substantially 20-40%, and its value is increased as the engine speed or the engine load increases. 
     Note that the swirl ratio of the swirl valve  18  of the engine of this embodiment is set slightly higher than 1.5:1 when its opening is 40%, and when the swirl valve  18  is fully closed (0%), the swirl ratio is increased to approximately 6:1. “Swirl ratio” is defined as a value obtained by dividing a value which is obtained from measuring an intake flow lateral angular speed for each valve lift and integrating the value, by an angular speed of a crankshaft. As described above, the opening of the swirl valve  18  is substantially controlled between 20 and 40% during the operation within the first warmed-up range A 1 . From this, in this embodiment, the opening of the swirl valve  18  within the first warmed-up range A 1  is set so that the swirl ratio inside the combustion chamber  6  becomes 1.5 or higher. 
     (b) Second Warmed-up Range 
     Within the second warmed-up range A 2 , similar to the first partially warmed-up range B 1 , the control for performing SPCCI combustion of the mixture gas is executed in the environment in which the air-fuel ratio inside the combustion chamber  6  is substantially the stoichiometric air-fuel ratio (λ=1). Since the control within the second warmed-up range A 2  is basically similar to the control described in (3-2(a)) (the control within the first partially warmed-up range B 1 ), its description is omitted here. 
     (c) Third Warmed-up Range 
     Within the third warmed-up range A 3 , similar to the second partially warmed-up range B 2 , the control for performing SPCCI combustion of the mixture gas is executed in the environment in which the air-fuel ratio inside the combustion chamber  6  is slightly richer than the stoichiometric air-fuel ratio (λ≤1). Since the control within the third warmed-up range A 3  is basically similar to the control described in (3-2(b)) (the control within the second partially warmed-up range B 2 ), its description is omitted here. 
     (d) Fourth Warmed-up Range 
     Within the fourth warmed-up range A 4 , similar to the third partially warmed-up range B 3 , the relatively traditional SI combustion is performed. Since the control within the fourth warmed-up range A 4  is basically similar to the control described in (3-2(c)) (the control within the third partially warmed-up range B 3 ), its description is omitted here. 
     (4) Setting of Open and Close timings of Intake Valve and Exhaust Valve within First Partially Warmed-up Range 
     The open and close timings of the intake valve  11  and the exhaust valve  12  set within the first partially warmed-up range B 1  (within an execution range of SPCCI combustion in a G/F lean environment) are described in detail. 
     (a) Open and Close Timings of Intake Valve 
       FIG. 13  is an operation map within the first partially warmed-up range B 1  divided into a plurality of sections based on the open and close timings of the intake valve  11 . 
     As illustrated in  FIG. 13 , the first partially warmed-up range B 1  is mainly divided into three partial ranges B 11  to B 13  based on the open and close timings of the intake valve  11 . When the three ranges are the first partial range B 11 , the second partial range B 12 , and the third partial range B 13 , the first partial range B 11  is a low load range where the engine load is below a given first load T 11 , the third partial range B 13  is a high load range where the engine load is higher than a given second load T 12 , and the second partial range B 12  is a medium load range which covers the remaining engine loads. 
     Here, the first partially warmed-up range B 1  described above is one example of “the operating range where the partial compression-ignition combustion in the G/F lean environment is performed,” the first load T 11  is one example of a “first reference load,” the first partial range B 11  is one example of a “first operating range,” and the second partial range B 12  is one example of a “second operating range.” Note that as described later, the first load T 11  is a boundary between a load range where the intake open timing IVO is advanced as the engine load increases and a load range where the intake open timing IVO is retarded as the engine load increases. 
       FIG. 14  is a chart of which a horizontal axis is the engine load and a vertical axis is the intake open timing IVO within the first partially warmed-up range B 1 , in which lines L 11 , L 12 , L 13 , and L 14  indicate the intake open timings IVO when the engine speed is the first speed N 1 , the second speed N 2 , the third speed N 3  and the fourth speed N 4 , respectively.  FIGS. 15A to 15D  illustrate the lines L 11  to L 14 , respectively. First to fourth speeds N 1  to N 4  here correspond to the N 1  to N 4  illustrated in  FIG. 13 , and the engine speed increases in this order. 
     Note that as illustrated in  FIG. 13 , etc., within a low engine speed range, a highest load (a highest value of the engine load) of the first partially warmed-up range B 1  is lower than the other ranges, and the line L 11  at which the engine speed is the first speed N 1  ends at a point at which the engine load is lower than the other lines L 12 , L 13 , and L 14 . 
     As illustrated in  FIGS. 15A to 15D , the intake open timing IVO is set on the advancing side than the TDC of the exhaust stroke within the entire first partially warmed-up range B 1 . 
     As illustrated in  FIGS. 15A to 15D , etc., within the entire first partial range B 11  in which the engine load is below the first load T 11  (at the respective engine speeds N 1  to N 4 ), the intake open timing IVO is set to advance as the engine load increases. In other words, within the first partial range B 11 , on the advancing side of TDC of the exhaust stroke, the intake open timing IVO is advanced as the engine load increases. In this embodiment, at each engine speed, the intake open timing IVO and the engine load have a substantially linear relationship, and the intake open timing IVO is advanced as the engine load increases. 
     Within the first partial range B 11 , at the highest engine load, i.e., the intake open timing IVO at the first load T 11  is set to a first timing IVO 1  which is near a most advanced timing thereof over all engine speeds. Within the first partial range B 11 , at each engine speed, the intake open timing IVO is changed between the first timing IVO 1  and a timing retarded therefrom by 20° C.A. 
     On the other hand, within the entire second partial range B 12  in which the engine load is higher than the first load T 11  and lower than the second load T 12 , the intake open timing IVO is set to be retarded as the engine load is higher (at all the engine speeds N 1  to N 4 ). In other words, within the second partial range B 12 , on the advancing side of TDC of the exhaust stroke, the intake open timing IVO is retarded as the engine load increases. 
     At a lowest engine load of the second partial range B 12 , that is, when the engine load is the first load T 11 , the intake open timing IVO is set to the first timing IVO 1 . Within the second partial range B 12 , the intake open timing IVO is retarded larger from the first timing IVO 1  as the engine load increases from the first load T 11 . Accordingly, when the engine load changes across the first load T 11 , the intake open timing IVO continuously changes. In other words, in this embodiment, the intake open timing IVO within the first partial range B 11  and the second partial range B 12  is set so that the intake open timing IVO continuously changes when the engine load changes across the first load T 11 . Also within the second partial range B 12 , the intake open timing IVO is changed within a range of substantially 20° C.A. 
     Within the second partial range B 12 , in an engine speed range from the first speed N 1  to the second speed N 2 , i.e., a low engine speed range, the engine load and the intake open timing IVO have a substantially linear relationship, and the intake open timing IVO is retarded as the engine load increases. On the other hand, within the second partial range B 12 , in an engine speed range from the second speed N 2  to the third speed N 3 , i.e., a high engine speed range, the change rate of the intake open timing IVO with respect to the engine load is lower as the engine load decreases, and when the engine load becomes high, the intake open timing IVO is changed relatively largely according to the engine load. 
     The intake open timing IVO in the third partial range B 13  in which the engine load is higher than the second load T 12  is set to be advanced as the engine load increases. 
     At each engine speed, the intake open timing IVO of when the engine load of the third partial range B 13  takes a lowest value, that is, when the engine load is the second load T 12  is set as same as that when the engine load of the second partial range B 12  takes a highest value. Within the third partial range B 13 , the intake open timing IVO is advanced as the engine load increases from the second load T 12 . Accordingly, when the engine load changes across the second load T 12 , the intake open timing IVO continuously changes. In other words, in this embodiment, the intake open timing IVO within the second partial range B 12  and the third partial range B 13  is set so that the intake open timing IVO continuously changes when the engine load changes across the second load T 12 . 
     Within the third partial range B 13 , at the first and third speeds N 1  and N 3 , the engine load and the intake open timing IVO have a substantially linear relationship, and the intake open timing IVO is advanced as the engine load increases. 
     On the other hand, in the third partial range B 13 , at the second speed N 2 , the change rate of the intake open timing IVO with respect to the engine load is set to be larger as the engine load decreases. More specifically, at the second speed N 2 , in a range where the engine load is higher than the boosting load T_t and the booster  33  performs the boost, the change rate of the intake open timing IVO with respect to the engine load is substantially 0, and in a range where the engine load is lower than the boosting load T_t and the booster  33  does not perform the boost, the intake open timing IVO is advanced as the engine load increases. 
     Moreover, within the third partial range B 13 , at the fourth speed N 4 , the change rate of the intake open timing IVO with respect to the engine load is set to be lower as the engine load decreases. 
       FIG. 16  is a chart illustrating a relationship between the engine load and the intake close timing IVC at each of the engine speeds N 1 , N 2 , N 3 , and N 4  of the first partially warmed-up range B 1 , corresponding to  FIG. 14 . As described above, the open and close timings of the intake valve  11  are changed while its open period is kept constant. Therefore, the intake close timing IVC is changed with respect to the engine load etc. similarly to the intake open timing IVO. 
     As illustrated in  FIG. 16 , the intake close timing IVC is retarded than BDC of the intake stroke in the entire first partially warmed-up range B 1 , and the intake valve  11  is closed during the intake stroke. Accordingly, within the first partially-warmed-up range B 1 , the intake close timing IVC is advanced to increase the amount of air introduced into the combustion chamber  6 . That is, when the intake valve  11  is opened during the intake stroke, air flows out from the combustion chamber  6  to the intake port  9  as the piston rises. Therefore, if the intake close timing IVC is advanced on the retarding side of BDC of the intake stroke, the amount of air flowing out to the intake port  9  is reduced and the amount of air confined in the combustion chamber  6  is increased. 
     (b) Close Timing of Exhaust Valve 
       FIG. 17  is an operation map of the first partially-warmed-up range B 1  divided based on the open and close timings of the exhaust valve  12 . 
     As illustrated in  FIG. 17 , the first partially warmed-up range B 1  is mainly divided into three partial ranges B 21  to B 23  based on the open and close timings of the exhaust valve  12 . When the three ranges are the fourth partial range B 21 , the fifth partial range B 22 , and the sixth partial range B 23 , the fourth partial range B 21  is a low load range where the engine load is below a given third load T 21 , the sixth partial range B 23  is a high load range where the engine load is higher than a given fourth load T 22 , and the fifth partial range B 22  is a medium load range which covers the remaining engine loads. 
     Here, the third load T 21  is one example of a “second reference load,” the fourth partial range B 21  is one example of a “third operating range,” and the fifth partial range B 22  is one example of a “fourth operating range.” Note that as described later, the third load T 21  is a boundary between a load range where the close timing EVC of the exhaust valve  12  is retarded as the engine load increases and a load range where the exhaust close timing EVC is advanced as the engine load increases. 
     The first load T 11  and the third load T 21  are set to be substantially the same value at the respective engine speeds and the first partial range B 11  and the fourth partial range B 21  are set in substantially the same range. For example, as illustrated in  FIG. 17 , the first load T 11  and the third load T 21  are set to the same value when the engine speed is high. On the other hand, the first load T 11  is set slightly higher than the third load T 21  when the engine speed is low. Note that this difference between the loads T 11  and T 21  is sufficiently small with respect to the full load (highest load) of the engine (e.g., less than 10% of the full load). 
       FIG. 18  is a chart of which a horizontal axis is the engine load and a vertical axis is the exhaust close timing EVC, in which lines L 21 , L 22 , L 23 , and L 24  indicate the exhaust close timing EVC when the engine speed is the first speed N 1 , the second speed N 2 , the third speed N 3  and the fourth speed N 4 , respectively.  FIGS. 19A to 19D  illustrate the lines L 21  to L 24 , respectively. 
     As illustrated in  FIG. 18 , etc., the exhaust close timing EVC is set on the retarding side of TDC of the exhaust stroke in the entire range of the first partially warmed-up range B 1 . That is, within the first partially warmed-up range B 1 , the exhaust valve  12  is closed at or after TDC of the exhaust stroke. 
     As illustrated in  FIGS. 19A to 19D , etc., within the fourth partial range B 21  in which the engine load is lower than the third load T 21 , the exhaust close timing EVC is set to retard as the engine load increases. In other words, within the fourth partial range B 21 , on the retarding side of TDC of the exhaust stroke, the exhaust close timing EVC is retarded as the engine load increases. In this embodiment, at each engine speed, the exhaust close timing EVC and the engine load have a substantially linear relationship, and the exhaust close timing EVC is advanced as the engine load increases. At a highest engine load of the fourth partial range B 21 , that is, when the engine load is the third load T 21 , the exhaust close timing EVC is set to a third timing EVC 1  for every engine speed. Within the fourth partial range B 21 , at each engine speed, the exhaust close timing EVC is changed between the third timing EVC 1  and a timing advanced therefrom by about 20° C.A. 
     On the other hand, the exhaust close timing EVC in the fifth partial range B 22  in which the engine load is higher than the third load T 21  and lower than the fourth load T 22  is set to advance as the engine load increases. In other words, within the fifth partial range B 22 , on the retarding side of TDC of the exhaust stroke, the exhaust close timing EVC is advanced as the engine load increases. 
     At a lowest engine load of the fifth partial range B 22 , that is, when the engine load is the third load T 21 , the exhaust close timing EVC is set to the third timing EVC 1 . Accordingly, when the engine load changes across the third load T 21 , the exhaust close timing EVC continuously changes. In other words, in this embodiment, the exhaust close timing EVC within the fourth partial range B 21  and the fifth partial range B 22  is set so that the exhaust close timing EVC continuously changes when the engine load changes across the third load T 21 . Also within the fifth partial range B 22 , the exhaust close timing EVC is changed within a range of substantially 20° C.A. 
     At a high engine speed side (fourth speed N 4 ) of the fifth partial range B 22 , the exhaust close timing EVC and the engine load have a substantially linear relationship and the exhaust close timing EVC is advanced as the engine load increases. On the other hand, at a low engine speed side (first to third speeds N 1  to N 3 ) of the fifth partial range B 22 , the exhaust close timing EVC and the engine load are kept substantially constant regardless of the engine load in a low engine load range, and is advanced as the engine load increases in a high engine load range. 
     The exhaust close timing EVC in the sixth partial range B 23  in which the engine load is higher than the fourth load T 22  is set to retard as the engine load increases. In other words, within the sixth partial range B 23 , on the retarding side of TDC of the exhaust stroke, the exhaust close timing EVC is retarded as the engine load increases. 
     At each engine speed, when the engine load of the sixth partial range B 23  takes a lowest value, that is, the exhaust close timing EVC when the engine load is the fourth load T 22  is set as same as that when the engine load of the fifth partial range B 22  takes a highest value. Accordingly, when the engine load changes across the fourth load T 22 , the exhaust close timing EVC continuously changes. In other words, in this embodiment, the exhaust close timing EVC within the fifth partial range B 22  and the sixth partial range B 23  is set so that the exhaust close timing EVC continuously changes when the engine load changes across the fourth load T 22 . 
     Note that as described above, the open and close timings of the exhaust valve  12  are changed while its open period is kept constant. Therefore, the open timing EVO of the exhaust valve  12  is changed with respect to the engine load, etc. similarly to the exhaust close timing EVC. 
     (c) Valve Overlap Period 
     As the open and close timings of the intake valve  11  and the exhaust valve  12  are set as described above, within the first partially warmed-up range B 1 , both the intake valve  11  and the exhaust valve  12  open across TDC of the exhaust stroke. Further, the first partially warmed-up range B 1  is divided as illustrated in  FIG. 20  based on a valve overlap period in which both of the intake valve  11  and the exhaust valve  12  are opened across TDC of the exhaust stroke (hereinafter, suitably referred to as a valve overlap period). Moreover, the relationship between the valve overlap period and the engine load at each of the engine speeds N 1  to N 4  (indicated by lines L 31  to L 34 , respectively) is as illustrated in  FIG. 21 . 
     The first partially warmed-up range B 1  is mainly divided into three partial ranges B 31  to B 33  based on the valve overlap period. When the three ranges are the O/L low load range B 31 , the O/L medium load range B 32 , and the O/L high load range B 33 , the O/L low load range B 31  is a low engine load range lower than a first O/L reference load T 31 , the O/L high load range B 33  is a high load range higher than a second O/L reference load T 32 , and the O/L medium load range B 32  is the remaining medium load range. 
     The first O/L reference load T 31  matches with the first load T 11 , and the O/L low load range B 31  matches with the first partial range B 11 . Further, the O/L low load range B 31  substantially matches with the fourth partial range B 21 . The second O/L reference load T 32  substantially matches with the second load T 12 , the O/L medium load range B 32  substantially matches with the second partial range B 12 , and the O/L high load range B 33  substantially matches with the third partial range B 13 . 
       FIG. 25  shows schematic charts illustrating valve lifts of the intake valve  11  and the exhaust valve  12  at each operation point when the engine speed is the second speed N 2 . The chart indicated by Y 1  in  FIG. 25  is for an operation point Y 1  included within the O/L low load range B 31 , the first partial range B 11  and the fourth partial range B 21 , and the chart indicated by Y 2  in  FIG. 25  is for an operation point Y 2  included within the O/L medium load range B 32 , the second partial range B 12  and the fifth partial range B 22 , and the chart indicated by Y 3  in  FIG. 25  is for an operation point Y 3  included within the O/L high load range B 33 , the third partial range B 13  and the sixth partial range B 23 . Note that the lowest chart of  FIG. 25  illustrates a change in the piston stroke (the position of the piston) with respect to the crank angle. 
     Within the entire (at all engine speeds of) O/L low load range B 31  (the first partial range B 11  and the fourth partial range B 21 ), the valve overlap period becomes longer as the engine load increases. In this embodiment, at each engine speed, the valve overlap period and the engine load have a substantially linear relationship, and the valve overlap period becomes longer as the engine load increases. 
     On the other hand, within the entire (at all engine speeds of) O/L medium load range B 32  (the second partial range B 12 ), the valve overlap period becomes longer than as the engine load increases. More specifically, at the third speed N 3 , the valve overlap period is kept constant regardless of the engine load on the low load side, and the valve overlap period is shortened as the engine load increases on the high load side. At other engine speeds N 1 , N 2 , and N 4 , the valve overlap period becomes shorter as the engine load increases within the entire O/L medium load range B 32 . 
     The overlap period when the engine load takes a lowest value within the O/L medium load range B 32  is set to the same value as the overlap period when the engine load takes a highest value within the O/L low load range B 31 , that is, when the engine load is the first O/L reference load T 31  (first load T 11 ). 
     Within the entire O/L high load range B 33  (at all engine speeds), the valve overlap period becomes longer than as the engine load increases. In this embodiment, the valve overlap period becomes longer as the engine load increases at each engine speed. 
     The overlap period when the engine load takes a lowest value within the O/L high load range B 33  is set to the same value as the overlap period when the engine load takes a highest value within the O/L medium load range B 32 , that is, when the engine load is the second O/L reference load T 32 . 
     (5) Setting of External EGR Ratio within First Partially Warmed-Up Range B 1   
       FIG. 22  is a chart of which a horizontal axis is the engine load and a vertical axis is the target external EGR ratio within the first partially warmed-up range B 1 , in which lines L 41 , L 42 , L 43 , and L 44  indicate the target external EGR ratios when the engine speed is the first speed N 1 , the second speed N 2 , the third speed N 3 , and the fourth speed N 4 , respectively.  FIGS. 23A to 23D  illustrate the lines L 41  to L 44 , respectively. 
     As illustrated in  FIGS. 22 and 23A to 23D , the first partially warmed-up range B 1  includes a segment in which the engine load is below the boosting load T_t, and substantially covers all the first partial range B 11 , the second partial range B 12  and the fourth partial range B 21 . The target external EGR ratio is substantially set higher as the engine load increases. Further, within a range where the engine load is higher than the boosting load T_t, the target external EGR ratio is set smaller as the engine load increases. 
     Specifically, when the engine load is below the first load T 11 , that is, within the first partial range B 11 , at all engine speeds N 1  to N 4 , the target external EGR ratio is increased as the engine load increases. 
     Within the range where the engine load is higher than the first load T 11  and lower than the second load T 12 , that is, within the second partial range B 12 , the target external EGR ratio is set for each engine speed as follows. 
     At the first speed N 1  of the second partial range B 12 , the target external EGR ratio is maintained substantially constant regardless of the engine speed. At the first speed N 1 , the target external EGR ratio at the lowest engine load of the second partial range B 12  and the target external EGR ratio at the highest engine load of the first partial range B 11  are set to the same value, and within the second partial range B 12 , the target external EGR ratio is maintained relatively high. 
     At the third speed N 3  of the second partial range B 12 , the target external EGR ratio is increased as the engine load increases in the low engine load range, and is decreased as the engine load increases in the high engine load range. Note that the reduction amount of the target external EGR ratio according to the engine load increase is extremely small, and at the third speed N 3  of the second partial range B 12 , the target external EGR ratio is substantially large when the engine load increases. 
     At the second speed N 2  and the fourth speed N 4  of the second partial range B 12 , the target external EGR ratio increases as the engine speed increases. 
     (6) About SI Ratio 
     As described above, in this embodiment SPCCI combustion combining SI combustion and CI combustion is performed within the first partially warmed-up range B 1 , etc. In this SPCCI combustion, it is important to control the ratio of SI combustion to CI combustion according to the operating condition. 
     In this embodiment, a SI ratio which is a ratio of a heat amount generated by SI combustion with respect to a total heat amount generated by SPCCI combustion (SI combustion and CI combustion) is used.  FIG. 7  is a chart illustrating this SI ratio and illustrating a change in heat generation rate (J/deg) according to the crank angle when SPCCI combustion occurs. A point X 1  in the waveform of  FIG. 7  is a heat generation point at which the heat generation rate rises with the start of SI combustion, and a crank angle θsi corresponding to this heat generation point X 1  is defined as the start timing of SI combustion. The point X 2  in the same waveform is a flection point appearing when the combustion mode switches from SI combustion to CI combustion and the crank angle θci corresponding to this flection point X 2  may be defined as the start timing of CI combustion. Further, an area R 1  of the waveform of the heat generation rate located on the advancing side of θci which is the start timing of CI combustion (from θsi to θci) is set as the heat generation amount by SI combustion, and an area R 2  of the waveform of the heat generation rate located on the retarding side of θci is set as a heat generation rate by CI combustion. Thus, the SI ratio defined by (heat generation amount by SI combustion)/(heat generation amount by SPCCI combustion) may be expressed by R 1 /(R 1 +R 2 ) using the respective areas R 1  and R 2 . That is, in this embodiment, the SI ratio=R 1 /(R 1 +R 2 ). 
     In CI combustion, since the mixture gas combusts a plurality of times simultaneously by self-ignition, a pressure increase rate easily increases compared to SI combustion which is caused by flame propagation. Therefore, especially if the SI ratio is carelessly decreased (that is, a proportion of CI combustion is increased) under a condition of high load and high fuel injection amount, loud noise is generated. On the other hand, since CI combustion does not occur unless the combustion chamber  6  is sufficiently heated and pressurized, under the condition that the load is low and the fuel injection amount is small, CI combustion is not started unless SI combustion has progressed to some extent, and the SI ratio naturally increases (that is, the proportion of CI combustion increases). In consideration of such circumstances, in this embodiment, the target SI ratio, which is the target value of SI ratio, is determined for each operating condition of the engine in the operation range where SPCCI combustion is performed. For example, within the first partially warmed-up range B 1 , the target SI ratio is set to decrease as the load substantially increases (that is, the proportion of CI combustion increases). Further correspondingly, in this embodiment, the target θci, which is the start timing of CI combustion when combustion conforming to the target SI ratio is performed, is determined for each operating condition of the engine. 
     In order to achieve the target SI ratio and the target θci described above, control amounts such as the timing of the main ignition by the spark plug  16 , the injection amount/injection timing of fuel from the injector  15 , the EGR ratio (the external EGR ratio and the internal EGR ratio) are adjusted for each operating condition. For example, as the ignition timing is advanced, a larger amount of fuel is combusted in SI combustion, and the SI ratio increases. Further, as the injection timing of the fuel is advanced, a larger amount of fuel is combusted in CI combustion, and the SI ratio decreases. Moreover, since a change in SI ratio is followed by a change in θci, changes in these control amounts (the main ignition timing, the injection timing, etc.) are elements for adjusting θci. 
     Based on the above tendency, in this embodiment, during the execution of SPCCI combustion, the main ignition timing, the fuel injection amount/injection timing, etc. are controlled in combination with each other to achieve the target SI ratio and target θci described above. 
     (7) Operations and Effects 
     As described above, in this embodiment, since the open and close timings of the intake valve  11 , the open and close timings of the exhaust valve  12 , and the target external EGR ratio are set as described above within the first partially warmed-up range B 1 , at each operation point of the first partially warmed-up range B 1 , the internal EGR gas, the external EGR gas, and air are suitably remained inside the combustion chamber  6  and the suitable SPCCI combustion is achieved. 
     For example, when a large amount of burnt gas remains inside the combustion chamber  6  when the temperature inside the combustion chamber  6  is difficult to rise due to the low engine load, the reaction of air and fuel is interrupted by the burnt gas and SI combustion easily becomes unstable. On the other hand, in this embodiment, within the first partial range B 11  (fourth partial range B 21 ), the intake open timing IVO is retarded on the advancing side of TDC of the exhaust stroke as the engine load decreases. That is, within the first partial range B 11 , the intake valve  11  is opened for a shorter period during the exhaust stroke as the engine load decreases. Therefore, within the first partial range B 11 , the amount of burnt gas (internal EGR gas) discharged from the combustion chamber  6  to the intake port  9  and flowing back into the combustion chamber  6  again is reduced, and it is avoided that the reaction of air and fuel is interrupted by the burnt gas, which improves the stability of SI combustion. If the stability of SI combustion is improved and SI combustion is suitably generated, the temperature inside the combustion chamber  6  near TDC of compression stroke is adjusted to the temperature at which CI combustion occurs (the temperature at which the mixture gas self-ignites), and extremely high thermal efficiency is achieved. Therefore, according to this embodiment, compared with conventional gasoline engines, drastic improvements in fuel efficiency and torque performance are achieved. 
     On the other hand, as the engine load increases, the temperature of burnt gas generated in the combustion chamber  6  increases. Therefore, when the engine load is high, even if the amount of burnt gas discharged to the intake port  9  and flowed back into the combustion chamber  6  again (internal EGR gas) is increased, the temperature inside the combustion chamber  6  is increased by the high-temperature burnt gas, flame propagation is promoted and the stability of SI combustion is improved. On the other hand, in this embodiment, as described above, within the first partial range B 11 , when the engine load is high, the intake open timing IVO is advanced on the advancing side of TDC of the exhaust stroke and the open period of the intake valve  11  is extended during the exhaust stroke. Therefore, within the first partial range B 11 , when the engine load is high, the stability of SI combustion is improved by increasing the amount of the high-temperature burnt gas remaining inside the combustion chamber  6  (internal EGR gas), and the suitable CI combustion is achieved. Further, increasing the amount of burnt gas as described above reduces a negative pressure in the intake passage, and therefore, a pumping loss is reduced and fuel efficiency is improved. 
     In this embodiment, within the second partial range B 12  in which the engine load is higher than the first partial range B 11 , that is, the range where the temperature of the burnt gas discharged to the intake port  9  and flowed back into the combustion chamber  6  again (internal EGR gas) further increases, the intake open timing IVO is retarded on the advancing side of TDC of the exhaust stroke as the engine load increases. Therefore, at the low load side of the second partial range B 12 , a relatively large amount of high-temperature burnt gas (internal EGR gas) is left inside the combustion chamber  6  and the stability of SI combustion is improved, while at the high load side, the high-temperature burnt gas (internal EGR gas) is prevented from being introduced into the combustion chamber  6  and an increase in combustion noise is avoided. That is, at the high load side of the second partial range B 12 , the temperature inside the combustion chamber  6  is prevented from excessively increasing and the start timing of CI combustion advancing earlier than a desired timing due to the large amount of high-temperature burnt gas (internal EGR gas), and the increase in combustion noise caused by advancing of the start timing of CI combustion is prevented. Here, the ignition timing needs to be retarded so that SI combustion occurs after TDC of compression stroke if combustion noise increases. However, retarding the ignition timing causes CI combustion at a timing when the piston descends significantly on the expansion stroke, decreasing fuel efficiency. Therefore, according to this embodiment, since the increase of combustion noise is avoided as described above, retarding the ignition timing and decreasing the fuel efficiency caused thereby are avoided. 
     Further in this embodiment, the intake open timing IVO and the exhaust close timing EVC within the partial ranges B 11  to B 23  are set so that they continuously change when the engine load changes across the first to fourth loads T 11  to T 22 . Therefore, when the engine load changes across the first to fourth loads T 11  to T 22 , the intake open timing IVO, the intake close timing IVC, the exhaust open timing EVO, and the exhaust close timing EVC are prevented from changing significantly, and the controllability of the intake and exhaust valves  11  and  12  is improved. In other words, the open and close timings of the intake valve  11  and the open and close timing of the exhaust valve  12  are prevented from significantly changing when the operation point of the engine shifts between adjacent partial ranges, and these open and close timings are reliably controlled to suitable timings. This is advantageous in achieving SPCCI combustion even if the operating state and the environmental condition change. 
     Further in this embodiment, the boost by the booster  33  is performed in a segment of the first partially warmed-up range B 1  in which the engine load is higher than the boosting load T_t. Therefore, within the range where the engine load is high, the amount of air introduced into the combustion chamber  6  is increased according to the engine load (fuel amount), and the air-fuel ratio of the mixture gas is more reliably brought near the stoichiometric air-fuel ratio. 
     Further, within the range in which the boost is performed in this manner, as illustrated in  FIG. 21 , since the valve overlap period is made longer as the engine load increases, the scavenging inside the combustion chamber  6  is promoted as the engine load increases so as to prevent excessively high-temperature burnt gas (internal EGR gas) from remaining inside the combustion chamber  6 . Therefore, the temperature inside the combustion chamber  6  is prevented from excessively increasing due to the large amount of high-temperature burnt gas (internal EGR gas). Moreover, at the high load side of the first partial range B 11  and the low load side of the second partial range B 12 , since the boost is stopped, the amount of burnt gas (internal EGR gas) remaining inside the combustion chamber  6  is secured and the combustion stability is improved more reliably. 
     Further, in this embodiment, within the range where the engine load including the second partial range B 12  is lower than the boosting load T_t, the external EGR ratio is substantially increased as the engine load increases. Therefore, within the second partial range B 12 , the amount of burnt gas (internal EGR gas) remaining inside the combustion chamber  6  when the engine load is high is reduced, while the external EGR gas introduced into the combustion chamber  6  is increased. Since the external EGR gas is introduced into the combustion chamber  6  after passing through the EGR passage  51 , the temperature of the external EGR gas at the point of it being introduced into the combustion chamber  6  is lower than the internal EGR gas remaining inside the combustion chamber  6 . Especially in this embodiment, the external EGR gas is introduced into the combustion chamber  6  after being cooled by the EGR cooler  52 . Therefore, the temperature of the external EGR gas introduced into the combustion chamber  6  is sufficiently lower than that of the internal EGR gas. By the above control of the internal and external EGR gas amounts, the temperature inside the combustion chamber  6  is prevented from being excessively increasing while the increase of the combustion noise is avoided. 
     Here, by advancing the exhaust close timing EVC on the retarding side of TDC of the exhaust stroke, the amount of the burnt gas discharged to the exhaust port  10  and introduced back into the combustion chamber  6  is reduced. In this regard, in this embodiment, within the third partial range B 21  (first partial range B 11 ), the exhaust close timing EVC is retarded on the retarding side of TDC of the exhaust stroke as the engine load increases. Therefore, within the third partial range B 21  (first partial range B 11 ) where the engine load is low and easily causes an unstable combustion, the amount of burnt gas remaining inside the combustion chamber  6  (internal EGR gas) is reduced particularly when the engine load is low, and the reaction between the fuel and air is promoted to improve the combustion stability. Moreover, when the engine load is relatively high and the temperature of the burnt gas (internal EGR gas) is relatively high, the amount of high-temperature burnt gas (internal EGR gas) is increased to improve the combustion stability. Additionally, within the fifth partial range A 5  where the temperature of the burnt gas (internal EGR gas) tends to increase, the amount of high-temperature burnt gas remaining inside the combustion chamber  6  (internal EGR gas) is reduced, and it is prevented that the temperature inside the combustion chamber  6  excessively increases and CI combustion starts excessively early. 
     Furthermore, in this embodiment, during the execution of SPCCI combustion (while operating within the first partially warmed-up range B 1 ), the timing of the main ignition by the spark plug  16 , etc. are adjusted so that the SI ratio being a value of a ratio of a heat amount generated by SI combustion with respect to a total heat generation amount in one combustion cycle matches with the target SI ratio determined according to the operating condition of the engine. Therefore, for example, the ratio of CI combustion is increased (i.e., the SI ratio is reduced) as much as possible within the extent that the combustion noise does not become excessive. This results in improving thermal efficiency by SPCCI combustion as much as possible. 
     (8) Modifications 
     In this embodiment, the SI ratio which is the ratio of the heat amount generated by SI combustion with respect to the total heat amount generated by SPCCI combustion is defined as R 1 /(R 1 +R 2 ) by using the areas R 1  and R 2  in the combustion waveform of  FIG. 7 , and the main ignition timing is adjusted to match the given target SI ratio; however, the SI ratio may be defined in other manners. 
     For example, SI ratio=R 1 /R 2  may be established. Further, the SI ratio may be defined using A 01  and A 02  illustrated in  FIG. 24 . That is, when the crank angle period of SI combustion (the combustion period on the advancing side of the flection point X 2 ) is Δθ1 and the crank angle period of CI combustion (the combustion period on the retarding side of the flection point X 2 ) is Δθ2, SI ratio=Δθ1/(Δθ1+Δθ2) or SI ratio=Δθ1/Δθ2 may be established. Alternatively, when a peak of the heat generation rate of SI combustion is ΔH1 and a peak of the heat generation rate of CI combustion is ΔH2, SI ratio=ΔH1/(ΔH1+ΔH2) or SI ratio=ΔH1/ΔH2 may be established. 
     It should be understood that the embodiments herein are illustrative and not restrictive, since the scope of the invention is defined by the appended claims rather than by the description preceding them, and all changes that fall within metes and bounds of the claims, or equivalence of such metes and bounds thereof, are therefore intended to be embraced by the claims. 
     DESCRIPTION OF REFERENCE CHARACTERS 
       2  Cylinder 
       11  Intake Valve 
       12  Exhaust Valve 
       13   a  Intake VVT (Intake Variable Mechanism) 
       14   a  Exhaust VVT (Exhaust Variable Mechanism) 
       15  Injector 
       16  Spark Plug 
       30  Intake Passage 
       32  Throttle Valve 
       33  Booster 
       40  Exhaust Passage 
       100  ECU (Controller)