Patent Publication Number: US-11028797-B2

Title: Engine control method and engine control device

Description:
TECHNICAL FIELD 
     The present disclosure relates to an engine control method and an engine control device. 
     BACKGROUND ART 
     Patent Document 1 discloses an engine that burns an air-fuel mixture in a combustion chamber by compression ignition in a predetermined low load and low speed range. The engine burns the air-fuel mixture by spark ignition in a region with higher load than that of a predetermined region and in a region with higher speed than that of the predetermined region. The engine further facilitates compression ignition of the air-fuel mixture by spark ignition using a spark plug near a compression top dead center also in the predetermined region. 
     Patent Document 2 discloses an engine that burns the air-fuel mixture in a fuel chamber by compression ignition in a high load range. In a high load and high speed range, this engine performs a small amount of fuel injection for ignition assistance at a post-stage of a compression stroke between preceding and succeeding injections for generating the air-fuel mixture for compression-ignition combustion. Accordingly, a rich air-fuel mixture is generated around a spark plug. Then, the spark plug ignites the rich air-fuel mixture to cause a flame, whereby the air-fuel mixture generated by the preceding injection is compressed and ignited near the compression top dead center. After that, the air-fuel mixture generated by the succeeding injection, which is performed at the same time as the compression ignition, is also compressed and ignited. 
     CITATION LIST 
     Patent Documents 
     Patent Document 1: Japanese Patent No. 4082292 
     Patent Document 2: Japanese Patent No. 5447435 
     SUMMARY OF THE INVENTION 
     Technical Problem 
     In an engine with a geometric compression ratio increased for the main purpose of improving thermal efficiency, an ignition unit performs spark ignition of an air-fuel mixture in a combustion chamber to cause combustion of the air-fuel mixture by flame propagation. Then, abnormal combustion including knocking may occur. For example, if the spark ignition is delayed to reduce the abnormal combustion, the combustion period becomes longer and the combustion center of gravity is largely away from compression top dead center. Accordingly, the thermal efficiency of the engine decreases. 
     The present disclosure increases the thermal efficiency of an engine. 
     Solution to the Problem 
     The present inventors have focused on utilizing what is called a “broken reaction zone.” In the broken reaction zone, a lean air-fuel mixture and/or a strong flow in the combustion chamber does not allow the progress of the combustion by flame propagation. Assume that the conditions inside of the combustion chamber fall within the broken reaction zone. When an air-fuel mixture is to be burned by the flame propagation, the made flame goes out. In a typical engine control, the air-fuel mixture is thus not spark-ignited, while the conditions inside the combustion chamber fall within the broken reaction zone. 
     However, the present inventors viewed the air-fuel mixture microscopically, newly finding the following. When an ignition unit ignites the air-fuel mixture in the broken reaction zone, a flame does not go out but is stored while being unable to cause the flame propagation. Once the conditions inside the combustion chamber fall out of the broken reaction zone, the stored flame starts to cause combustion of the air-fuel mixture at once. Based on the finding that the stored flame starts to cause combustion of the air-fuel mixture at once, once the conditions inside the combustion chamber fall out of the broken reaction zone, the present inventors completed the disclosed technique related to a new combustion mode. 
     Specifically, the present disclosure relates to an engine control method that executes a cycle including an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke in a combustion chamber. The engine control method includes a fuel supply step and an ignition step. In the fuel supply step, the fuel supply unit supplies fuel into the combustion chamber. In the ignition step, an ignition unit arranged in the combustion chamber makes a flame after the supply of the fuel into the combustion chamber and at a timing when a flow strength in the combustion chamber is greater than a predetermined value in the compression stroke during or before a post-mid stage where the compression stroke is divided into four stages of a pre-stage, a pre-mid stage, a post-mid stage, and a post-stage. 
     According to the above configuration, in the fuel supply step, the fuel is supplied into the combustion chamber to generate an air-fuel mixture in the combustion chamber. In the fuel supply step, the fuel may be supplied into the combustion chamber before the ignition step which will be described later. For example, when the fuel is directly injected into the combustion chamber by the fuel supply unit, the fuel may be injected into the combustion chamber in a period from the intake stroke to an initial stage of the compression stroke or in a period from the intake stroke to a pre-mid stage of the compression stroke. This allows for supply of the fuel into the combustion chamber before the ignition step. The “initial stage of the compression stroke” may be that initial stage where the compression stroke is evenly divided into three stages of an initial stage, a middle stage, and a final stage. On the other hand, for example, if the fuel supply unit is configured to inject the fuel into an intake port(s) connected to the combustion chamber (and into the combustion chamber), the fuel may be injected into the intake port(s) before the intake stroke (including the exhaust stroke). This allows for introduction of the fuel together with intake air into the combustion chamber in the period of the intake stroke and supply of the fuel into the combustion chamber before the ignition step. 
     In the compression stroke after the fuel supply step, the ignition step is executed. In the ignition step, the ignition unit makes the flame in the combustion chamber. The ignition unit may be, for example, a spark plug that causes a spark discharge between electrodes. Further, the ignition unit may be configured to cause an arc discharge or a plasma discharge, for example. By applying energy to the air-fuel mixture, the ignition unit makes the flame in the combustion chamber. 
     In the period of the intake stroke, an intake flow is generated in the combustion chamber by introducing intake air through intake ports into the combustion chamber. The generated intake flow weakens once near intake bottom dead center. However, the flow inside the combustion chamber gradually strengthens in the period from the initial stage to the middle stage of the compression stroke when a piston moves toward top dead center, due to what is called a “spin-up phenomenon.” After that, the flow inside the combustion chamber gradually weakens to the post-stage of the compression stroke. In the ignition step, the flame is made in the combustion chamber at the post-mid stage of a compression stroke or before the post-mid stage of the compression stroke at a timing when the flow strength in the combustion chamber is equal to or higher than a predetermined value. As the flow strength in the combustion chamber is high, even when the flame is made, it is possible to keep the flame as-is in the combustion chamber without causing combustion by flame propagation. That is, the ignition step is performed when the conditions inside the combustion chamber fall within the broken reaction zone. Here, in the ignition step, the ignition unit may perform a plurality of discharges. This increases the number of flames made in the combustion chamber and allows for relatively strong flow in the combustion chamber, thereby making it possible to diffuse the flames made into the combustion chamber. 
     After the ignition step, the flow strength in the combustion chamber decreases as the piston comes closer to compression top dead center. After the post-stage with the flow strength in the combustion chamber decreased, the conditions in the combustion chamber fall out of the broken reaction zone. Further, the temperature and the pressure in the combustion chamber increase at the post-stage of the compression stroke due to motoring. The flame made in the ignition step and stored in the combustion chamber starts the combustion of the air-fuel mixture at the post-stage of the compression stroke or in the expansion stroke. More specifically, the combustion of the air-fuel mixture starts at once by autoignition near compression top dead center. The center of gravity of this combustion is close to compression top dead center, which improves the thermal efficiency of the engine. In addition, this combustion mode requires a shorter combustion period and thus reduces knocking. 
     The ignition unit may make the flame in the combustion chamber at a timing when a tumble ratio in the combustion chamber is equal to or higher than a predetermined value in the compression stroke. 
     The intake port is configured as a so-called tumble port. Accordingly, when a tumble flow (i.e., vertical vortex) is generated in the combustion chamber in the intake stroke, it is possible to increase the strength of the tumble flow in the combustion chamber (i.e., to increase the tumble ratio as the index indicating the strength of the tumble flow) due to the spin-up phenomenon in a period from the initial stage to the middle stage of the compression stroke. The ignition unit makes the flame in the combustion chamber at a timing when a tumble ratio in the combustion chamber is equal to or higher than a predetermined value. Accordingly, it is possible to store the flame in the combustion chamber without allowing the progress of the combustion by flame propagation. 
     The ignition unit may make the flame in the combustion chamber in a compression stroke with the engine having a speed equal to or higher than a predetermined value. 
     When the speed of the engine is high, the flow in the combustion chamber strengthens. In the compression stroke with the engine having a speed equal to or higher than a predetermined value, the ignition unit makes flame in the combustion chamber. Accordingly, it is possible to store the flame in the combustion chamber without allowing the progress of the combustion by flame propagation. 
     The fuel supply unit may supply the fuel in the combustion chamber at a timing when the ignition unit makes the flame in the combustion chamber so that an air-fuel mixture is generated at least around the ignition unit, the air-fuel mixture having an air-fuel mass ratio A/F or a gas-fuel mass ratio G/F, in which gas includes air, higher than a stoichiometric air-fuel ratio. 
     The broken reaction zone relates to the two parameters of the fuel concentration of the mixture and the flow strength in the combustion chamber. When the fuel concentration of the air-fuel mixture is low, the conditions inside the combustion chamber fall within the broken reaction zone which does not allows for progression of combustion by flame propagation. If the air-fuel mixture having a lean A/F or G/F with respect to the stoichiometric air-fuel ratio is generated at least around the ignition unit in the fuel supply step, it is possible to store flame in the combustion chamber without allowing for the progress of combustion by flame propagation when making the flame in the combustion chamber. The air-fuel mixture is generated at a relatively early phase in the combustion chamber. However, the air-fuel mixture is a lean air-fuel mixture with a low fuel concentration, thereby making it possible to reduce the risk of pre-ignition. 
     In the engine control method, it is possible to perform, after the ignition step, a second fuel supply step in which the fuel supply unit supplies the fuel into the combustion chamber to increase a fuel concentration of an air-fuel mixture in the combustion chamber. 
     In the second fuel supply step, additional fuel supply by the fuel supply unit leads to a high fuel concentration of the air-fuel mixture. As the fuel concentration of the air-fuel mixture increases, the conditions inside the combustion chamber fall out of the broken reaction zone. After the second fuel supply step, autoignition of the flame stored in the combustion chamber occurs to start the combustion of the air-fuel mixture at once. The adjustment of the timing of the additional fuel supply and/or the adjustment of the fuel amount to be additionally supplied make it possible to adjust the timing at which the combustion of the air-fuel mixture starts. 
     The engine may have a geometric compression ratio of 14 or more. The engine control method disclosed herein improves thermal efficiency, while reducing abnormal combustion in an engine with a high compression ratio. 
     The present disclosure also relates to an engine control device. The engine control device includes: a combustion chamber that executes a cycle including an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke; an ignition unit arranged in the combustion chamber; and a fuel supply unit configured to supply fuel into the combustion chamber, in which the ignition unit makes a flame after the supply of the fuel into the combustion chamber by the fuel supply unit and at a timing when a flow strength in the combustion chamber is greater than a predetermined value in the compression stroke during or before a post-mid stage where the compression stroke is divided into four stage of a pre-stage, a pre-mid stage, a post-mid stage, and a post-stage. 
     The ignition unit may make the flame in the combustion chamber at a timing when a tumble ratio in the combustion chamber is equal to or higher than a predetermined value in the compression stroke. 
     The ignition unit may make the flame in the combustion chamber in the compression stroke with the engine having a speed equal to or higher than a predetermined value. 
     The fuel supply unit may supply the fuel in the combustion chamber at a timing when the ignition unit makes the flame in the combustion chamber so that an air-fuel mixture is generated at least around the ignition unit, the air-fuel mixture having an air-fuel mass ratio A/F or a gas-fuel mass ratio G/F, in which gas includes air, higher than a stoichiometric air-fuel ratio. 
     The fuel supply unit may be configured to supply, after the ignition unit has made the flame, the fuel into the combustion chamber to increase a fuel concentration of the air-fuel mixture in the combustion chamber. 
     The engine may have a geometric compression ratio of 14 or more. 
     Advantages of the Invention 
     The engine control method and device described above increase thermal efficiency of an engine. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  illustrates a configuration of an engine system. 
         FIG. 2  illustrates a configuration of a combustion chamber. 
         FIG. 3  is a top view illustrating configurations of the combustion chamber and an intake system. 
         FIG. 4  is a block diagram illustrating a configuration of an engine control device. 
         FIG. 5  illustrates operating range maps of the engine shown in  FIG. 1 . 
         FIG. 6  illustrates fuel injection times, ignition times, and combustion waveforms in respective operating states. 
         FIG. 7  illustrates a rig tester for swirl ratio measurement. 
         FIG. 8  illustrates a relationship between the opening rate of a secondary passage and the swirl ratio. 
         FIG. 9  illustrates a change in the flow strength in the combustion chamber from an intake stroke to a compression stroke. 
       The upper illustration of  FIG. 10  shows an example change in an ignition time with respect to an engine speed, whereas the lower illustration of  FIG. 10  shows an example change in the start of second injection with respect to the engine speed. 
         FIG. 11  is a flowchart related to the control of fuel injection and an ignition time. 
         FIG. 12  illustrates a configuration of an engine system different from that in  FIG. 1 . 
         FIG. 13  illustrates an operating range map of the engine shown in  FIG. 12 . 
     
    
    
     DESCRIPTION OF EMBODIMENT 
     An exemplary embodiment of an engine control device and an engine control method will now be described in detail with reference to the drawings.  FIG. 1  illustrates a configuration of an engine system including an engine  1 .  FIG. 2  illustrates a configuration of a combustion chamber  17 . In this  FIG. 2 , the upper illustration corresponds to a top view of the combustion chamber  17 , whereas the lower illustration is a cross-sectional view taken along the line II-II of the upper illustration.  FIG. 3  illustrates configurations of the combustion chamber  17  and an intake system.  FIG. 4  is a block diagram illustrating a configuration of an engine control device. In  FIG. 1 , the intake side of the engine is located on the left of the drawing plane, whereas the exhaust side on the right of the drawing plane. In  FIGS. 2 and 3 , the intake side is located on the right of the drawing plane, whereas the exhaust side on the left of the drawing plane. 
     The engine  1  is a four-stroke engine that operates while repeating a cycle including an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke in the combustion chamber  17 . The engine  1  is mounted in a four-wheeled motor vehicle. The motor vehicle travels in accordance with the operation of the engine  1 . Fuel of the engine  1  is gasoline in this exemplary configuration. The fuel may be gasoline containing bioethanol, for example. The fuel for the engine  1  may be any fuel as long as the fuel is liquid fuel which contains at least gasoline. 
     &lt;Configuration of Engine&gt; 
     The engine  1  is a multi-cylinder engine. As shown in  FIG. 1 , the engine  1  includes an engine body  2  with the combustion chamber  17 . The engine body  2  includes a cylinder block  12  and a cylinder head  13  above the cylinder block  12 . Inside the cylinder block  12 , a plurality of cylinders  11  are arranged. Note that  FIGS. 1 and 2  each show only one of the cylinders  11 . 
     A piston  3  is slidably fitted into each of the cylinders  11 . The piston  3  is coupled to a crankshaft  15  via a connecting rod  14 . The piston  3  defines the combustion chamber  17 , together with each of the cylinders  11  and the cylinder head  13 . Note that the “combustion chamber” meant here is not limited to the space defined when the piston  3  reaches a compression top dead center. The term “combustion chamber” may be used in a broader sense. That is, the “combustion chamber” may be the space defined by the piston  3 , the cylinder  11 , and the cylinder head  13 , regardless of the position of the piston  3 . The expression “in the combustion chamber” and “in the cylinder” may be used in substantially the same meaning. 
     As shown in the lower illustration of  FIG. 2 , the lower surface of the cylinder head  13 , that is, the ceiling of the combustion chamber  17 , includes inclined surfaces  1311  and  1312 . The inclined surface  1311  is an upward slope extending from the intake side toward an injection axis X 2  of an injector  6 , which will be described later. On the other hand, the inclined surface  1312  is an upward slope extending from the exhaust side toward the injection axis X 2 . The ceiling of the combustion chamber  17  is in the shape of what is called a “pent roof”. 
     The upper surface of the piston  3  protrudes toward the ceiling of the combustion chamber  17 . Defined above the upper surface of the piston  3  is a cavity  31 . The cavity  31  is recessed from the upper surface of the piston  3 . The cavity  31  faces the injector  6  which will be described later. The center of the cavity  31  is shifted toward the exhaust side with respect to a center axis X 1  of the cylinder  11  but agrees with the injection axis X 2  of the injector  6 . 
     The cavity  31  includes a projection  311 . The projection  311  is located on the injection axis X 2  of the injector  6 . This projection  311  is in the shape of a substantial cone extending upward from the bottom of the cavity  31  toward the ceiling of the combustion chamber  17 . The cavity  31  is symmetric about the injection axis X 2  of the injector  6 . 
     The cavity  31  also includes a recess  312  around the projection  311 . The recess  312  surrounds the entire circumference of the projection  311 . The circumferential side surface of the recess  312  is inclined with respect to the injection axis X 2  from the bottom of the cavity  31  toward the opening of the cavity  31 . The inner diameter of the cavity  31  of the recess  312  gradually increases from the bottom of the cavity  31  toward the opening of the cavity  31 . 
     Note that the shape of the combustion chamber  17  is not limited to that illustrated in  FIG. 2 . That is, the shapes of the cavity  31 , the upper surface of the piston  3 , and the ceiling of the combustion chamber  17 , for example, may be changed as appropriate. For example, the cavity  31  may be symmetric about the center axis X 1  of the cylinder  11 . The inclined surfaces  1311  and  1312  may be symmetric about the center axis X 1  of the cylinder  11 . The cavity  31  may have a shallow bottom that is shallower than the bottom of the recess  312  so as to face a spark plug  25 , which will be described later. 
     The engine  1  has a geometric compression ratio ranging from 14 to 30. As will be described later, the engine  1  performs, in some operating ranges, SPark Controlled Compression Ignition (SPCCI) combustion that is a combination of spark ignition (SI) combustion and compression ignition (CI) combustion. The SI combustion is accompanied by flame propagation started by forcible ignition of an air-fuel mixture in the combustion chamber  17 . The CI combustion is started by autoignition of the air-fuel mixture in the combustion chamber  17 . The combined mode of these SI and CI combustions is as follows. The air-fuel mixture in the combustion chamber  17  is forcibly ignited to start the combustion by the flame propagation. Then, the heat generation and the flame propagation in the SI combustion increase the pressure, which leads to the compression ignition for burning unburned air-fuel mixture in the combustion chamber  17 . In this engine  1 , there is no need to significantly increase the temperature of the combustion chamber  17  when the piston  3  reaches compression top dead center, that is, the compression end temperature, for the autoignition of the air-fuel mixture. 
     The cylinder head  13  includes two intake ports  18  for each cylinder  11 . As shown in  FIG. 3 , the intake ports  18  include two intake ports of first and second intake ports  181  and  182 . The first and second intake ports  181  and  182  are aligned along the axis of the crankshaft  15 , that is, in the front-rear direction of the engine body  2 . The intake ports  18  communicate with the combustion chamber  17 . Although not shown in detail, the intake ports  18  are what are called “tumble ports”. That is, each intake port  18  is in a shape causing a tumble flow in the combustion chamber  17  in the intake stroke. 
     Each intake port  18  includes an intake valve  21 . The intake valve  21  opens and closes the intake port  18  between the combustion chamber  17  and the intake port  18 . The engine  1  includes a valve train mechanism for the intake valves  21 . The intake valves  21  are opened and closed by a valve train mechanism at predetermined timing. The valve train mechanism for the intake valves  21  may be a variable valve train mechanism allowing variable valve timing and/or variable valve lift. 
     In this exemplary configuration, the variable valve train mechanism includes, as shown in  FIG. 4 , an intake electric sequential-valve timing (S-VT)  23 . The intake electric S-VT  23  is a variable valve train mechanism of a phase type that causes the intake valves  21  to open at a constant angle and to open and close at variable times. The intake electric S-VT  23  continuously changes the rotational phase of an exhaust camshaft within a predetermined angular range. Accordingly, the opening and closing times of the intake valve  21  change continuously. Note that the valve train mechanism for the intake valves  21  may include a hydraulic S-VT instead of the electric S-VT. The valve train mechanism for the intake valves  21  may include a variable valve train mechanism that changes the amounts of lift of the intake valves  21  and/or a variable valve train mechanism that changes the opening angles (or the opening periods) of the intake valves  21 . 
     The cylinder head  13  also includes two exhaust ports  19  for each cylinder  11 . As shown in  FIG. 3 , the exhaust ports  19  include two exhaust ports, i.e., a first exhaust port  191  and a second exhaust port  192 . The first and second exhaust ports  191  and  192  are aligned in the front-rear direction of the engine body  2 . The exhaust ports  19  communicate with the combustion chamber  17 . 
     Each exhaust port  19  includes an exhaust valve  22 . The exhaust valve  22  opens and closes the exhaust port  19  between the combustion chamber  17  and the exhaust port  19 . The engine  1  includes a valve train mechanism for the exhaust valves  22 . The exhaust valves  22  are opened and closed by the valve train mechanism at predetermined timing. The valve train mechanism for the exhaust valves  22  may be a variable valve train mechanism allowing variable valve timing and/or variable valve lift. 
     In this exemplary configuration, the variable valve train mechanism includes, as shown in  FIG. 4 , an exhaust electric S-VT  24 . The exhaust electric S-VT  24  is a variable valve train mechanism of a phase type that causes the exhaust valves  22  to open at a constant angle and to open and close at variable times. The exhaust electric S-VT  24  continuously changes the rotational phase of an exhaust camshaft within a predetermined angular range. Accordingly, the opening and closing times of the exhaust valves  22  change continuously. Note that the valve train mechanism for the exhaust valves  22  may include a hydraulic S-VT instead of the electric S-VT. The valve train mechanism for the exhaust valves  22  may include a variable valve train that changes the amounts of lift of the exhaust valves  22  and/or a variable valve train mechanism that changes the opening angles (or the opening periods) of the exhaust valves  22 . 
     In the engine  1 , the intake and exhaust electric S-VTs  23  and  24  adjust the length of the overlap period between the opening times of the intake valves  21  and the closing times of the exhaust valves  22 . Accordingly, hot burned gas is confined in the combustion chamber  17 . That is, internal exhaust gas recirculation (EGR) gas is introduced into the combustion chamber  17 . The adjustment of the length of the overlap period allows scavenging of the residual gas (burned gas) in the combustion chamber  17 . 
     The cylinder head  13  includes the injector  6  for each cylinder  11 . The injector  6  directly injects the fuel into the combustion chamber  17 . The injector  6  is an example a fuel supply unit. The injector  6  faces the inside of the combustion chamber  17  at the valley of the pent roof, at which the inclined surface  1311  on the intake side and the inclined surface  1312  on the exhaust side intersect each other, and is opposed to the cavity  31 . 
     As shown in  FIG. 2 , the injection axis X 2  of the injector  6  is parallel to the center axis X 1  of the cylinder  11  and is closer to the exhaust side than the center axis X 1  of the cylinder  11 . The injection axis X 2  of the injector  6  agrees with the position of the projection  311  of the cavity  31 . Note that the injection axis X 2  of the injector  6  may agree with the center axis X 1  of the cylinder  11 . In this case as well, it is desirable that the injection axis X 2  of the injector  6  agree with the position of the projection  311  of the cavity  31 . 
     Although not shown in detail, the injector  6  includes a multi-port combustion injection valve with a plurality of nozzle ports. As indicated by the two-dot chain lines in  FIG. 2 , the injector  6  injects the fuel so that the fuel spray spreads radially from the center of the combustion chamber  17  and spreads obliquely downward from the ceiling of the combustion chamber  17 . 
     In this exemplary configuration, the injector  6  has ten nozzle ports. The nozzle ports are arranged at equal angles along the circumference of the injector  6 . As shown in the upper illustration of  FIG. 2 , the axes of the nozzle ports are shifted along the circumference of the injector  6  with respect to the spark plug  25  which will be described later. That is, the spark plug  25  is interposed between the axes of two adjacent nozzle ports. This arrangement reduces the risk of direct contact of the spray of the fuel injected from the injector  6  with the spark plug  25 , thereby making it possible to avoid making any electrode wet. 
     A fuel supply system  61  is connected to the injector  6 . The fuel supply system  61  includes a fuel tank  63  configured to store the fuel, and a fuel supply passage  62  that connects the fuel tank  63  and the injector  6  together. The fuel supply passage  62  includes a fuel pump  65  and a common rail  64 . The fuel pump  65  pumps out the fuel to the common rail  64 . 
     In this exemplary configuration, the fuel pump  65  is a plunger pump driven by the crankshaft  15 . The common rail  64  stores the fuel pumped out from the fuel pump  65  at a high fuel pressure. When the injector  6  opens, the fuel stored in the common rail  64  is injected from a nozzle port of the injector  6  into the combustion chamber  17 . 
     The fuel supply system  61  can supply fuel to the injector  6  at a high pressure of 30 MPa or more. The maximum fuel pressure of the fuel supply system  61  may be about 120 MPa, for example. The pressure of the fuel to be supplied to the injector  6  may vary in accordance with the operating state of the engine  1 . Note that the configuration of the fuel supply system  61  is not limited to the configuration described above. 
     The cylinder head  13  includes the spark plug  25  attached to each of the cylinders  11 . The spark plug  25  performs a spark discharge between electrodes arranged in the combustion chamber  17 , thereby forcibly igniting the air-fuel mixture in the combustion chamber  17 . The spark plug  25  is an example of an ignition unit. 
     In this exemplary configuration, as shown in  FIG. 2 , the spark plug  25  is closer to the intake side with respect to the center axis X 1  of the cylinder  11  in the combustion chamber  17 . The spark plug  25  is adjacent to the injector  6  and interposed between the two intake ports. In addition, the spark plug  25  is attached to the cylinder head  13 , while being inclined such that the bottom of the plug is closer to the center of the combustion chamber  17  than the top of the plug. The electrodes of the spark plug  25  face the combustion chamber  17  and are located near the ceiling of the combustion chamber  17 . 
     One side surface of the engine body  2  is connected to an intake passage  40 . The intake passage  40  communicates with the intake ports  18  of each cylinder  11  and with the combustion chamber  17  via the intake ports  18 . Through the intake passage  40 , the gas introduced into the combustion chamber  17  flows. Located at the upstream end of the intake passage  40  is an air cleaner  41  that filters fresh air. Located near the downstream end of the intake passage  40  is a surge tank  42 . A part of the intake passage  40  downstream of the surge tank  42  forms independent passages that branch off for the respective cylinders  11 . The downstream end of each independent passage is connected to the intake ports  18  of the associated one of the cylinders  11 . 
     A throttle valve  43  is interposed between the air cleaner  41  and the surge tank  42  in the intake passage  40 . The opening degree of the throttle valve  43  is adjusted to the amount of fresh air to be introduced into the combustion chamber  17 . 
     In the intake passage  40 , a supercharger  44  is provided downstream of the throttle valve  43 . The supercharger  44  supercharges the gas inside the intake passage  40  to be introduced into the combustion chamber  17 . 
     In the exemplary configuration, the supercharger  44  is a mechanical supercharger driven by the engine body  2 . The mechanical supercharger  44  may be of a Roots type, for example. The mechanical supercharger  44  may have any configuration. The mechanical supercharger  44  may be of a Lysholm type, a vane type, or a centrifugal type. 
     An electromagnetic clutch  45  is interposed between the supercharger  44  and the engine body  2 . The electromagnetic clutch  45  transmits a driving force from the engine body  2  to the supercharger  44  or blocks the driving force between the supercharger  44  and the engine body  2 . As will be described later, the supercharger  44  is turned on and off by an engine control unit (ECU)  10  that determines whether to engage or disengage the electromagnetic clutch  45 . Accordingly, the engine  1  determines whether or not to supercharge the gas to be introduced into the combustion chamber  17  by the supercharger  44 . 
     In the intake passage  40 , the intercooler  46  is provided downstream of the supercharger  44 . The intercooler  46  cools the gas compressed by the supercharger  44 . The intercooler  46  may be of a water-cooling type, for example. Alternatively, the intercooler  46  may be of an oil cooling type. 
     The intake passage  40  is also connected to a bypass passage  47 . The bypass passage  47  connects the upstream part of the supercharger  44  and the downstream part of the intercooler  46  in the intake passage  40  together so as to bypass the supercharger  44  and the intercooler  46 . The bypass passage  47  includes an air bypass valve  48 . The air bypass valve  48  adjusts the flow rate of gas flowing through the bypass passage  47 . 
     When the supercharger  44  is turned off, that is, when the electromagnetic clutch  45  is disengaged, the air bypass valve  48  fully opens. Accordingly, the gas flowing in the intake passage  40  bypasses the supercharger  44 , that is, passes through none of the supercharger  44  or the intercooler  46  but through the bypass passage  47  and flows into the surge tank  42 . The gas is then introduced into the combustion chamber  17  of the engine  1 . The engine  1  operates without supercharging, that is, with natural aspiration. 
     When the supercharger  44  is turned on, that is, when the electromagnetic clutch  45  is engaged, the gas flowing through the intake passage  40  passes through the supercharger  44  and the intercooler  46  and then flows into the surge tank  42 . At this time, if the air bypass valve  48  is open, a portion of the gas that has passed through the supercharger  44  flows back from the surge tank  42  through the bypass passage  47  to the upstream side of the supercharger  44 . The back flow rate of such gas varies depending on the opening degree of the air bypass valve  48 . The supercharging pressure of the gas inside the intake passage  40  may be controlled by adjusting the opening degree of the air bypass valve  48 . 
     In this exemplary configuration, the supercharger  44 , the bypass passage  47 , and the air bypass valve  48  constitute a supercharging system  49  in the intake passage  40 . 
     The engine  1  includes a swirl generating unit that generates a swirl flow in the combustion chamber  17 . As shown in  FIG. 3 , the swirl generating unit is a swirl control valve  56  attached to the intake passage  40 . The swirl control valve  56  is located in a secondary passage  402  out of primary and secondary passages  401  and  402  that communicate with the first and second intake ports  181  and  182 , respectively. 
     The swirl control valve  56  is an opening degree adjustment valve capable of narrowing the cross-section of the secondary passage  402 . In the combustion chamber  17 , a swirl flow occurs which has a strength corresponding to the opening degree of the swirl control valve  56 . The swirl flow circulates counterclockwise in  FIG. 3  as indicated by the arrows (see the white arrows in  FIG. 2  as well). 
     At a lower opening degree of the swirl control valve  56 , the flow rate of the intake air flowing from the first intake port  181  into the combustion chamber  17  relatively increases, whereas the flow rate of the intake air flowing from the second intake port  182  into the combustion chamber  17  relatively decreases, out of the first and second intake ports  181  and  182  aligned in the front-rear direction of the engine body  2 . This causes a stronger swirl flow in the combustion chamber  17 . When the swirl control valve  56  opens at a higher degree, the flow rates of the intake air flowing into the combustion chamber  17  from the first and second intake ports  181  and  182  are substantially equal to each other. This causes a weaker swirl flow in the combustion chamber  17 . When the swirl control valve  56  fully opens, the swirl flow does not occur. 
     Instead of or in addition to attaching the swirl control valve  56  to the intake passage  40 , the swirl generating section may employ the following configuration. The opening periods of two intake valves  21  to allow introduction of the intake air from only one of the intake valves  21  into the combustion chamber  17 . With the opening of only one of the two intake valves  21 , the intake air is unevenly introduced into the combustion chamber  17 , which allows generation of a swirl flow in the combustion chamber  17 . In addition, each intake port  18  may have an innovative shape so that the swirl generating unit generates a swirl flow in the combustion chamber  17 . 
     The other side surface of the engine body  2  is connected to an exhaust passage  50 . The exhaust passage  50  communicates with the exhaust ports  19  of each cylinder  11  and with the combustion chamber  17  via the exhaust ports  19 . Through the exhaust passage  50 , exhaust gas discharged from the combustion chamber  17  flows. Although not shown in detail, an upstream part of the exhaust passage  50  forms independent passages that branch off for the respective cylinders  11 . The upstream end of each independent passage is connected to the exhaust ports  19  of associated one of the cylinders  11 . 
     The exhaust passage  50  is provided with an exhaust gas purification systems having a plurality of (two in the example shown in  FIG. 1 ) catalyst converters. Although not shown, an upstream catalyst converter is located inside an engine compartment. This upstream catalyst converter includes a three-way catalyst  511  and a gasoline particulate filter (GPF)  512 . On the other hand, a downstream catalyst converter is located outside the engine compartment. This downstream catalyst converter includes a three-way catalyst  513 . 
     Note that the configuration of the exhaust gas purification system is not limited to the exemplary configuration shown in the figure. For example, the GPF  512  may be omitted. The catalyst converters are not limited to the three-way catalysts  511  and  513 . The order of the three-way catalysts  511  and  513  and the GPF  512  may be changed as appropriate. 
     An EGR passage  52  constituting an external EGR system is interposed between the intake passage  40  and the exhaust passage  50 . The EGR passage  52  is for returning a portion of the burned gas into the intake passage  40  and connects the intake passage  40  and the exhaust passage  50  together. The upstream end of the EGR passage  52  is connected between the upstream and downstream catalyst converters in the exhaust passage  50 . On the other hand, a downstream end of the EGR passage  52  is connected to the upstream side of the supercharger  44  in the intake passage  40 . The external EGR system is what is called a “low-pressure EGR system”. 
     The EGR passage  52  includes a water-cooling EGR cooler  53 . The EGR cooler  53  cools the burned gas. The EGR passage  52  also includes an EGR valve  54 . The EGR valve  54  adjusts the flow rate of the burned gas flowing through the EGR passage  52 . The backflow rate of the cooled burned gas, that is, external EGR gas, may be adjusted by changing the opening degree of the EGR valve  54 . 
     In this exemplary configuration, an EGR system  55  includes the external EGR system including the EGR passage  52  and the EGR valve  54 , and the internal EGR system including the intake and exhaust electric S-VTs  23  and  24  described above. 
     The engine system includes the ECU  10  for operating the engine  1 . The ECU  10  is a controller including a known microcomputer as a base element. As shown in  FIG. 4 , the ECU  10  includes a central processing unit (CPU)  101 , memory  102  such as random-access memory (RAM) and read-only memory (ROM), and an input and output (I/O) bus  103 . The CPU  101  executes programs. The memory  102  stores the programs and data. The I/O bus  103  receives and outputs electrical signals. 
     This ECU  10  is connected to the injectors  6  described above, the spark plugs  25 , the intake electric S-VT  23 , the exhaust electric S-VT  24 , the fuel supply system  61 , the throttle valve  43 , the EGR valve  54 , the electromagnetic clutch  45  of the supercharger  44 , the air bypass valve  48 , and the swirl control valve  56 . As shown in  FIGS. 1 and 4 , the ECU  10  is also connected to various types of sensors SW 1  to SW 16 . The sensors SW 1  to SW 16  output detection signals to the ECU  10 . 
     The sensors includes the following. An airflow sensor SW 1  and a first intake air temperature sensor SW 2  are arranged downstream of the air cleaner  41  in the intake passage  40 . A first pressure sensor SW 3  is located downstream of the part of the intake passage  40  connected to the EGR passage  52  and upstream of the supercharger  44 . A second intake air temperature sensor SW 4  is located downstream of the supercharger  44  in the intake passage  40  and upstream of the part of the intake passage  40  connected to the bypass passage  47 . A second pressure sensor SW 5  is attached to the surge tank  42 . A pressure indicating sensor SW 6  is attached to the cylinder head  13  in association with each cylinder  11 . An exhaust gas temperature sensor SW 7  is located in the exhaust passage  50 . 
     The airflow sensor SW 1  detects the flow rate of the fresh air flowing through the intake passage  40 . The first intake air temperature sensor SW 2  detects the temperature of the fresh air flowing through the intake passage  40 . The first pressure sensor SW 3  detects the pressure of the gas flowing into the supercharger  44 . The second intake air temperature sensor SW 4  detects the temperature of the gas flowing out from the supercharger  44 . The second pressure sensor SW 5  detects the pressure of the gas downstream of the supercharger  44 . The indicator sensor SW 6  detects the pressures inside the combustion chamber  17 . The exhaust gas temperature sensor SW 7  detects the temperature of the exhaust gas discharged from the combustion chamber  17 . 
     The sensors further include the following. A linear O 2  sensor SW 8  is disposed upstream of the upstream catalyst converter in the exhaust passage  50 . A lambda O 2  sensor SW 9  is disposed downstream of the three-way catalyst  511  in the upstream converter. A water temperature sensor SW 10 , a crank angle sensor SW 11 , an intake cam angle sensor SW 12 , and an exhaust cam angle sensor SW 13  are attached to the engine body  2 . An accelerator position sensor SW 14  is attached to an accelerator pedal mechanism. An EGR differential pressure sensor SW 15  is disposed in the EGR passage  52 . A fuel pressure sensor SW 16  is attached to the common rail  64  of the fuel supply system  61 . 
     The linear O 2  sensor SW 8  and the lambda O 2  sensor SW 9  each detect the oxygen concentration in the exhaust gas. The water temperature sensor SW 10  detects the temperature of the coolant. The crank angle sensor SW 11  detects the rotation angle of the crankshaft  15 . The intake cam angle sensor SW 12  detects the rotation angle of the intake camshaft. The exhaust cam angle sensor SW 13  detects the rotation angle of the exhaust camshaft. The accelerator position sensor SW 14  detects the accelerator position. The EGR differential pressure sensor SW 15  detects the differential pressure between the upstream and downstream sides of the EGR valve  54 . The fuel pressure sensor SW 16  detects the pressure of the fuel to be supplied to the injectors  6 . 
     Based on the detection signals of these sensors, the ECU  10  determines the operating state of the engine  1  and calculates the control amounts of the devices. The ECU  10  outputs control signals related to the calculated control amounts to the injectors  6 , the spark plugs  25 , the intake electric S-VT  23 , the exhaust electric S-VT  24 , the fuel supply system  61 , the throttle valve  43 , the EGR valve  54 , the electromagnetic clutch  45  of the supercharger  44 , the air bypass valve  48 , and the swirl control valve  56 . 
     For example, the ECU  10  sets the target torque of the engine  1  and determines the target supercharging pressure based on the detection signal of the accelerator position sensor SW 12  and a map set in advance. Then, the ECU  10  adjusts the opening degree of the air bypass valve  48  based on the target supercharging pressure and the differential pressure before and after the supercharger  44  obtained from the detection signals of the first pressure sensor SW 3  and the second pressure sensor SW 5 . Accordingly, feedback control is performed so that the supercharging pressure reaches the target supercharging pressure. 
     The ECU  10  sets the target EGR rate, that is, the ratio of the EGR gas to the entire gas in the combustion chamber  17 , based on the operating state of the engine  1  and the map set in advance. Then, the ECU  10  determines the target amount of EGR gas based on the target EGR rate and the amount of the intake air based on the detection signal of the accelerator position sensor SW 12 . The ECU  10  adjusts the opening degree of the EGR valve  54  based on the differential pressure before and after the EGR valve  54  obtained from the detection signal of the EGR differential pressure sensor SW 15 . Through the determination and the adjustment, the ECU  10  performs feedback control so that the amount of the external EGR gas to be introduced into the combustion chamber  17  reaches the target amount of EGR gas. 
     The ECU  10  further executes feedback control of the air-fuel ratio upon satisfaction of predetermined control conditions. Specifically, the ECU  10  adjusts the amount of fuel injection by the injectors  6  based on the oxygen concentration in the exhaust gas detected by the linear O 2  sensor SW 8  and the lambda O 2  sensor SW 9  so that the air-fuel ratio of the air-fuel mixture reaches a desired value. 
     Details of the control of the engine  1  by the ECU  10  will be described later. 
     &lt;Operating Range Map of Engine&gt; 
       FIG. 5  illustrates operating range maps  501  and  502  of the warmed-up engine  1 . The operating range maps  501  and  502  of the engine  1  are defined by the load and speed of the engine  1 , and are divided into two ranges based on the magnitude of the speed of the engine  1 . 
     Specifically, the two ranges are: an SPCCI range ( 1 ) at a lower speed, specifically, at an engine speed lower than N 1 ; and a CI range ( 2 ) at a higher speed, specifically, at an engine speed higher than or equal to N 1 . The SPCCI range ( 1 ) may here include low and medium speeds, if the entire operating range of the engine  1  is divided into three of low, medium, and high speed ranges in the direction of the speed. On the other hand, the CI range ( 2 ) may include the high speed range. The speed N 1  may be about 4,000 rpm, for example. 
     In  FIG. 5 , for easier understanding, each of the operating range maps  501  and  502  of the engine  1  is divided into two ranges. The map  501  shows the conditions of the air-fuel mixture and the combustion modes in operating states  601  to  604  of the engine  1  and driving and non-driving ranges of the supercharger  44 . The map  502  shows the opening degree of the swirl control valve  56  in each range. Note that the two-dot chain lines in  FIG. 5  represent road-load lines of the engine  1 . 
     The engine  1  performs combustion by compressed autoignition for the main purpose of improving fuel efficiency and the exhaust gas performance. More specifically, the engine  1  performs the SPCCI combustion described above in the SPCCI range ( 1 ). In the CI range ( 2 ), the engine  1  performs the CI combustion. Now, the operation of the engine  1  in the operating states  601  to  604  shown in  FIG. 5  will be described in detail with reference to the fuel injection and ignition times shown in  FIG. 6 . In  FIG. 6 , the horizontal axis indicates the crank angle which advances from the left to the right in the drawing plane of  FIG. 6 . 
     &lt;Engine Control in Low-Load Operation in SPCCI Range ( 1 )&gt; 
     While operating in the SPCCI range ( 1 ), the engine  1  performs the SPCCI combustion as described above. In the combustion by the autoignition, the autoignition timing largely changes with a variation in the temperature in each combustion chamber  17  before the start of the compression. In the SPCCI combustion, the spark plug  25  forcibly ignites the air-fuel mixture in the combustion chamber  17  to cause the SI combustion of the air-fuel mixture by the flame propagation. The heat generated in the SI combustion increases the temperature in the combustion chamber  17 . The increase in the temperature in the combustion chamber  17  by the flame propagation causes the CI combustion of the unburned mixture by the autoignition. By adjusting the amount of heat generated by the SI combustion, the variation in the temperature is compensated in the combustion chamber  17  before the start of the compression. That is, even if the temperature in the combustion chamber  17  varies before the start of the compression, the autoignition timing can be controlled by adjusting the start of the SI combustion through the adjustment of the ignition timing, for example. 
     In  FIG. 6 , the reference character  601  denotes an example including fuel injection times (reference characters  6011  and  6012 ), an ignition time (reference character  6013 ), and a combustion waveform (reference character  6014 ) in a low load operating mode  601  of the engine  1  in the SPCCI range ( 1 ). The combustion waveform represents a change in the heat generation rate with respect to the crank angle. 
     In the SPCCI combustion, the spark plug  25  ignites the air-fuel mixture at a predetermined timing near compression top dead center (TDC on the right of  FIG. 6 ). Accordingly, the combustion by the flame propagation starts. The heat generation is more moderate in the SI combustion than in the CI combustion. The waveform of the heat generation rate has thus a relatively shallow slope at the rising. Although not shown, the pressure fluctuation (dp/dθ) in the combustion chamber  17  is also more moderate in the SI combustion than in the CI combustion. 
     Once the SI combustion increases the temperature and the pressure inside the combustion chamber  17 , the autoignition of the unburned air-fuel mixture occurs. In the example of  FIG. 6 , at the autoignition timing, the waveform of the heat generation rate changes from the shallower slope to a steeper slope (see the reference character  6014 ). That is, the waveform of the heat generation rate has an inflection point at the start of the CI combustion. 
     After the start of the CI combustion, the SI and CI combustions are performed in parallel. Since the CI combustion generates more heat than the SI combustion and thus has a relatively high heat generation rate. However, since the CI combustion is performed after the compression top dead center, the piston  3  is lowered by the motoring, which does not allow the CI combustion to cause an excessively steep slope of the waveform of the heat generation rate. In addition, the pressure fluctuation (dp/dθ) in the CI combustion becomes relatively moderate. 
     The pressure fluctuation (dp/dθ) may be used as an index representing combustion noise. The SPCCI combustion can reduce the pressure fluctuation (dp/dθ) as described above, thereby making it possible to avoid causing too much combustion noise. This allows for suppression of combustion noise to an acceptable level or lower. 
     The SPCCI combustion ends with an end of the CI combustion. The CI combustion requires a shorter combustion period than the SI combustion. Thus, the combustion ends earlier in the SPCCI combustion than in the SI combustion. In other words, in the SPCCI combustion, the combustion end can be closer to compression top dead center in the expansion stroke. Therefore, the SPCCI combustion is more advantageous in improving the fuel efficiency of the engine  1  than the SI combustion. 
     The EGR system  55  introduces the EGR gas into the combustion chamber  17  at a low load of the engine  1  in the SPCCI range ( 1 ) to improve the fuel efficiency of the engine  1 . 
     Specifically, a positive overlap period, in which both the intake and exhaust valves  21  and  22  are open, is provided near the exhaust top dead center. This leads to performing the internal EGR in which a part of the exhaust gas discharged from the inside of the combustion chamber  17  to the intake and exhaust ports  18  and  19  returns and is reintroduced into the combustion chamber  17 . The internal EGR introduces hot burned gas (i.e., internal EGR gas) into the combustion chamber  17  and thus increases the temperature in the combustion chamber  17 , which is advantageous in stabilizing the SPCCI combustion. 
     At a low load of the engine  1 , the EGR valve  54  is fully closed. The external EGR gas is not introduced into the combustion chamber  17 . 
     The supercharger  44  is turned off at a low load of the engine  1  in the SPCCI range ( 1 ). Specifically, the supercharger  44  is turned off (see S/C OFF) at a lower speed and at low and medium loads in the SPCCI range ( 1 ). Even at the low and medium loads of the engine  1 , the supercharger  44  is turned on (see S/C ON) at a higher speed of the engine  1  and increases the supercharging pressure to ensure a required filling amount of the intake air. 
     When the supercharger  44  is turned off not to supercharge the gas in the intake passage  40 , the pressure inside the intake passage  40  is relatively low. The internal EGR gas is thus introduced into the combustion chamber  17  in the positive overlap period as described above. 
     When the supercharger  44  is turned on to supercharge the gas in the intake passage  40 , the pressure inside the intake passage  40  is relatively high. The gas in the intake passage  40  thus passes through the combustion chamber  17  of the engine body  2  and blows to the exhaust passage  50  in the positive overlap period. Accordingly, the burned gas remaining in the combustion chamber  17  is pushed out to the exhaust passage  50  and scavenged. 
     In an operation of the engine  1  in the SPCCI range ( 1 ), the swirl control valve  56  is fully closed or at a predetermined closing angle. Accordingly, a relatively strong swirl flow occurs in the combustion chamber  17 . The swirl flow is stronger on the periphery of the combustion chamber  17  and weaker at the center. As described above, each intake port  18  is the tumble port. Thus, an oblique swirl flow with a tumble component and a swirl component occurs in the combustion chamber  17 . 
     At a low load of the engine  1 , the swirl ratio is 4 or more, for example. The swirl ratio is here defined as follows. The “swirl ratio” is the value obtained through dividing, by the angular velocity of the engine, the value obtained through measuring and integrating the lateral angular velocities of the intake flows for the respective valve lifts. The lateral angular velocity of the intake flow may be obtained based on measurement using a rig tester shown in  FIG. 7 . 
     The tester shown in  FIG. 7  is configured as follows. The cylinder head  13  is placed upside down on a base to connect the intake ports  18  to an intake air supplier (not shown). Meanwhile, a cylinder  36  is placed on the cylinder head  13  to connect an impulse meter  38  including a honeycomb rotor  37  at the upper end of the cylinder  36 . The lower surface of the impulse meter  38  is located at a distance of 1.75D from the mating surface between the cylinder head  13  and the cylinder block. The term “D” means here the diameter of the cylinder bore. The tester measures, using the impulse meter  38 , the torque acting on the honeycomb rotor  37  due to the swirl flow (see the arrow in  FIG. 7 ) generated in the cylinder  36  in accordance with the supply of the intake air. Based on the torque thus measured, the lateral angular velocity of the intake flow is obtained. 
       FIG. 8  shows a relationship between the opening degree of the swirl control valve  56  and the swirl ratio of the engine  1 . In  FIG. 8 , the opening degree of the swirl control valve  56  is expressed by the opening rate of the secondary passage  402  with respect to its fully opened cross-section. When the swirl control valve  56  is fully closed, the opening rate of the secondary passage  402  is 0%. With an increase in the opening degree of the swirl control valve  56 , the opening rate of the secondary passage  402  becomes larger than 0%. When the swirl control valve  56  is fully open, the opening rate of the secondary passage  402  is 100%. 
     As illustrated in  FIG. 8 , when the swirl control valve  56  is fully closed, the swirl ratio of the engine  1  is about 6. At a low load of the engine  1  in the SPCCI range ( 1 ), the swirl ratio may range from 4 to 6, both inclusive. The opening degree of the swirl control valve  56  may be adjusted to the opening rate ranging from 0% to 15%, both inclusive. 
     The air-fuel ratio (A/F) of the air-fuel mixture is higher than the stoichiometric air-fuel ratio in the entire combustion chamber  17  at a low load of the engine  1  in the SPCCI range ( 1 ). That is, the excessive air ratio λ of the air-fuel mixture in the entire combustion chamber  17  is more than 1 (λ&gt;1). More specifically, the A/F of the air-fuel mixture is 30 or more in the entire combustion chamber  17 . This ratio allows for reduction in raw NOx and improvement in the exhaust gas performance. 
     In the SPCCI range ( 1 ), in the low load operating mode  601  of the engine  1 , the air-fuel mixture is stratified between the center and periphery of the combustion chamber  17 . At the center of the combustion chamber  17 , the spark plug  25  is disposed. The periphery of the combustion chamber  17  is around the center and in contact with the liner of the cylinder  11 . The center of the combustion chamber  17  may be defined as a region with a weaker swirl flow, whereas the periphery of the combustion chamber  17  may be defined as a region with a stronger swirl flow. 
     The fuel concentration of the air-fuel mixture at the center of the combustion chamber  17  is higher than that on the periphery of the combustion chamber  17 . Specifically, the A/F of the air-fuel mixture at the center of the combustion chamber  17  ranges from 20 to 30, whereas the A/F of the air-fuel mixture on the periphery of the combustion chamber  17  is 35 or more. The air-fuel ratio is the value as of ignition, which may also apply to the following description. 
     At a low load of the engine  1  in the SPCCI range ( 1 ), the injector  6  injects the fuel into the combustion chamber  17  a plurality of times in the compression stroke (see the reference characters  6011  and  6012  in  FIG. 6 ). Specifically, the injector injects the fuel at the middle stage and final stage of the compression stroke. The mid-stage and post-stage of the compression stroke may be the middle stage and final stage where the compression stroke is evenly divided into three stages of an initial stage, a middle stage, and a final stage with respect to the crank angle. 
     The fuel injected at the middle stage of the compression stroke diffuses inside the combustion chamber  17  by the ignition time to generate the air-fuel mixture at the center and on the periphery of the combustion chamber  17 . The fuel is injected at the post-stage of the compression stroke, that is, in a short time until the ignition, and is thus transported by the swirl flow to the vicinity of the spark plug  25  at the center of the combustion chamber  17  without being diffused much. The fuel forms, together with a portion of the fuel injected at the middle stage of the compression stroke, the air-fuel mixture at the center of the combustion chamber  17 . As described above, the air-fuel mixture is stratified at the center and on the periphery of the combustion chamber  17 . 
     After the end of the fuel injection, the spark plug  25  ignites the air-fuel mixture at the center of the combustion chamber  17  at predetermined timing before compression top dead center (see the reference character  6013 ). At this time, the air-fuel mixture contains the fuel at a relatively high concentration at the center of the combustion chamber  17 , which improves the ignitability and stabilizes the SI combustion by the flame propagation. The stabilization of the SI combustion allows for the start of the CI combustion at an appropriate timing. That is, in the SPCCI combustion, the controllability of the CI combustion improves. As a result, at a low load of the engine  1  in the SPCCI range ( 1 ), it is possible to achieve both the reduction in combustion noise, and the improvement in fuel efficiency due to a shorter combustion period. 
     &lt;Engine Control in Medium-Load Operation in SPCCI Range ( 1 )&gt; 
     In  FIG. 6 , the reference character  602  denotes an example including fuel injection times (reference characters  6021  and  6022 ), an ignition time (reference character  6023 ), and a combustion waveform (reference character  6024 ), in a medium load operation of the engine  1  in the SPCCI range ( 1 ). 
     The EGR system  55  introduces the EGR gas into each combustion chamber  17  in a medium load operation of the engine  1  as in a low load operation. Specifically, in a low load and a lower speed operation of the engine  1  within the medium load range, the positive overlap period, in which both the intake and exhaust valves  21  and  22  are open, is provided near the exhaust top dead center. This leads to the internal EGR in which a part of the exhaust gas discharged from the inside of the combustion chamber  17  to the intake and exhaust ports  18  and  19  returns and is reintroduced into the combustion chamber  17 . That is, the internal EGR gas is introduced into the combustion chamber  17 . 
     In a high load or a high speed operation of the engine  1  within the medium load range, the supercharger  44  is turned on to ensure a filling amount of the intake air required with an increase in the amount of fuel injection. When the supercharger  44  is turned on to supercharge the gas in the intake passage  40 , the pressure inside the intake passage  40  is relatively high. The residual gas (i.e., hot burned gas) in the combustion chamber  17  is thus scavenged in the positive overlap period as described above. 
     In a medium load operation of the engine  1 , external EGR is performed in which the exhaust gas cooled by the EGR cooler  53  is introduced through the EGR passage  52  into the combustion chamber  17 . That is, the external EGR gas with a lower temperature than that of the internal EGR gas is introduced into the combustion chamber  17 . The introduction of at least one of the internal EGR gas and the external EGR gas into the combustion chamber  17  leads to adjustment of the temperature in the combustion chamber  17  to an appropriate temperature. Note that the EGR rate increases with an increase in the load on the engine  1 . 
     In a medium load operation of the engine  1 , the swirl control valve  56  is fully closed or at a predetermined closing angle as in a low load operation. Accordingly, a strong swirl flow with a swirl ratio of 4 or more occurs in the combustion chamber  17 . With an increase in the strength of the swirl flow, the turbulent energy inside the combustion chamber  17  increases, which causes rapid propagation of the flame in the SI combustion and stabilization of the SI combustion. The stabilization of the SI combustion increases the controllability of the CI combustion. This causes an appropriate timing of the CI combustion in the SPCCI combustion. As a result, combustion noise decreases and fuel efficiency improves. In addition, the variation in the torque among the cycles decreases. 
     The air-fuel ratio (A/F) of the air-fuel mixture is equal to the stoichiometric air-fuel ratio (i.e., A/F=14.7) in the entire combustion chamber  17  in a medium load operation of the engine  1 . At the stoichiometric air-fuel ratio, a three-way catalyst purifies the exhaust gas discharged from the combustion chamber  17  so that the engine  1  has an excellent exhaust gas performance. The A/F of the air-fuel mixture may fall within the purification window of the three-way catalyst. Therefore, the excessive air ratio λ of the air-fuel mixture may be 1.0±0.2. 
     In a medium load operation of the engine  1 , the injector  6  injects the fuel into the combustion chamber  17  separately in the intake and compression strokes (see the reference characters  6021  and  6022  in  FIG. 6 ). Specifically, the injector performs first injection  6021  to inject the fuel in the period from the middle stage to final stage of the intake stroke, and second injection  6022  to inject the fuel at a second half of the compression stroke. The middle stage and final stage of the intake stroke may be the middle stage and final stage where the intake stroke is evenly divided into three stages of an initial stage, a middle stage, and a final stage with respect to the crank angle. On the other hand, first and second halves of the compression stroke may be the first and second halves of the compression stroke when the compression stroke is divided into two of first and second halves with respect to the crank angle. 
     In the first injection  6021 , the fuel is injected at a time away from the ignition time. At the time of the injection, the piston  3  is away from the top dead center, the fuel reaches a squish area  171  outside the cavity  31  and is distributed substantially evenly in the combustion chamber  17  to generate air-fuel mixture. In the second injection  6022 , the fuel is injected when the piston  3  is closer to compression top dead center and thus enters the cavity  31  to generate air-fuel mixture in the area inside of the cavity  31 . 
     In accordance with the second injection  6022  for injecting the fuel into the cavity  31 , the gas flow occurs in the area inside the cavity  31 . If a longer time is required to the ignition, the turbulent energy in the combustion chamber  17  weakens with the progress of the compression stroke. However, the second injection  6022  is closer to the ignition time than the first injection  6021 , which allows the spark plug  25  to ignite the air-fuel mixture in the area inside the cavity  31  while maintaining the large turbulent energy in the cavity  31 . Accordingly, the speed of the SI combustion increases. With an increase in the combustion rate of the SI combustion, the SI combustion stabilizes, thereby leading to improvement in the controllability of the CI combustion by the SI combustion. 
     In the second injection  6022  at the second half of the compression stroke, latent heat of fuel vaporization reduces the temperature in the combustion chamber  17  to reduce induction of abnormal combustion such as pre-ignition or knocking. The fuel injected in the second injection  6022  can be stably burned by the flame propagation. The ratio between the amount of the first injection  6021  and the amount of the second injection  6022  may be 95:5, for example. 
     In the combustion chamber  17 , a substantially homogeneous air-fuel mixture with an excessive air ratio λ of 1.0±0.2 as a whole is generated by the injector  6  performing the first and second injections  6021  and  6022 . Since the air-fuel mixture is substantially homogeneous, fuel efficiency improves with a decrease in unburned loss and the exhaust gas performance improves with a decrease in smoke (soot). 
     The spark plug  25  ignites the air-fuel mixture at a predetermined timing before compression top dead center (see the reference character  6023 ), thereby causing the combustion of the air-fuel mixture by flame propagation. After the combustion has started by the flame propagation, the autoignition of the unburned air-fuel mixture is performed, initiating the CI combustion. The fuel injected in the second injection  6022  is mainly subjected to the SI combustion. The fuel injected in the first injection  6021  is mainly subjected to the CI combustion. 
     &lt;Engine Control in High Load Operation in SPCCI Range ( 1 )&gt; 
     In  FIG. 6 , reference character  603  denotes an example including a fuel injection time (reference character  6031 ), an ignition time (reference character  6032 ), and a combustion waveform (reference character  6033 ) in a high load operation of the engine  1  within the SPCCI range ( 1 ). 
     The EGR system  55  also introduces the EGR gas into each combustion chamber  17  in a high load operation of the engine  1 . 
     Specifically, in a high load operation of the engine  1 , the external EGR is performed in which the exhaust gas cooled by the EGR cooler  53  is introduced through the EGR passage  52  into the combustion chamber  17 . The EGR rate continuously increases with an increase in the load on the engine  1  within the medium and high load ranges of in the SPCCI range ( 1 ). The introduction of the external EGR gas cooled by the EGR cooler  53  into the combustion chamber  17  leads to adjustment of the temperature in the combustion chamber  17  to an appropriate temperature and reduction in induction of abnormal combustion, such as pre-ignition or knocking, of the air-fuel mixture. 
     At a load on the engine  1  closer to the full load, there is a need to increase the amount of fresh air to be introduced into the combustion chamber  17  to cope with an increase in the amount of fuel. For this purpose, at a load of the engine  1  closer to the full load in the SPCCI range ( 1 ), the EGR rate of the external EGR may decrease. 
     In a high load operation of the engine  1  as well, the positive overlap period, in which both the intake and exhaust valves  21  and  22  are open, is provided near the exhaust top dead center. 
     In a high load operation of the engine  1  as well, the supercharger  44  is turned on (see S/C ON) throughout the high load range to increase the supercharging pressure. This allows for scavenge of the residual gas (i.e., the burned gas) in the combustion chamber  17  in the positive overlap period. 
     In a high load operation of the engine  1  as well, the swirl control valve  56  is fully closed or at a predetermined closing angle. Accordingly, a strong swirl flow with a swirl ratio of 4 or more occurs in the combustion chamber  17 . 
     In a high load operation of the engine  1 , the air-fuel ratio (A/F) of the air-fuel mixture is lower than or equal to the stoichiometric air-fuel ratio (i.e., the excessive air ratio of the air-fuel mixture is expressed by λ≤1) in the entire combustion chamber  17 . 
     In a high load operation mode  603  of the engine  1 , the injector  6  starts injecting the fuel in the intake stroke (see the reference character  6031  in  FIG. 6 ). Specifically, the fuel injection  6031  may start injecting the fuel at 280° CA before compression top dead center. The fuel injection  6031  may continue over the intake stroke and end in the compression stroke. The fuel injection  6031  may start at the first half of the intake stroke, which allows the fuel spray to hit the opening edge of the cavity  31 . A part of the fuel enters the squish area  171  of the combustion chamber  17 , that is, the area outside the cavity  31  (see  FIG. 2 ), whereas the rest enters the area inside the cavity  31 . At this time, the swirl flow is stronger on the periphery of the combustion chamber  17 , and weaker at the center of the combustion chamber  17 . Accordingly, the fuel that has entered the area inside the cavity  31  enters more inward than the swirl flow. 
     The fuel that has entered the swirl flow remains in the swirl flow from the intake stroke to the compression stroke and forms the air-fuel mixture for the CI combustion on the periphery of the combustion chamber  17 . The fuel that has entered the inside of the swirl flow also remains inside the swirl flow from the intake stroke to the compression stroke and forms the air-fuel mixture for the SI combustion at the center of the combustion chamber  17 . 
     In a high load operation of the engine  1 , the fuel concentration of the air-fuel mixture is set to be higher on the periphery of the combustion chamber  17  than at the center. In addition, the amount of fuel in the air-fuel mixture is set to be larger on the periphery of the combustion chamber  17  than at the center. 
     Specifically, the excessive air ratio λ of the air-fuel mixture is 1 or lower in one preferred embodiment at the center of the combustion chamber  17 , and 1 or lower and, in one preferred embodiment, lower than 1 on the periphery of the combustion chamber  17 . At the center of the combustion chamber  17 , the air-fuel ratio (A/F) of the air-fuel mixture may range from 13 to the stoichiometric air-fuel ratio (i.e., 14.7), for example. Alternatively, at the center of the combustion chamber  17 , the air-fuel ratio of the air-fuel mixture may be higher than the stoichiometric air-fuel ratio. 
     On the periphery of the combustion chamber  17 , the air-fuel ratio of the air-fuel mixture may range from 11 to the stoichiometric air-fuel ratio, for example, and may range from 11 to 12 in one preferred embodiment. The excessive air ratio λ less than 1 on the periphery of the combustion chamber  17  increases the amount of the fuel in the air-fuel mixture on the periphery, whereby the latent heat of fuel vaporization decreases the temperature. In the combustion chamber  17 , the air-fuel ratio of the air-fuel mixture may range from 12.5 to the stoichiometric air-fuel ratio, for example, and may range from 12.5 to 13 in one preferred embodiment. 
     Near compression top dead center, the spark plug  25  ignites the air-fuel mixture in the combustion chamber  17  (see the reference character  6032 ). The spark plug  25  may perform the ignition, for example, after compression top dead center. Since the spark plug  25  is located at the center of the combustion chamber  17 , the ignition of the spark plug  25  starts the SI combustion of the air-fuel mixture by the flame propagation at the center. Since the fuel concentration of the air-fuel mixture is higher around the spark plug  25 , the flame stably propagates after the ignition of the spark plug  25  in the SPCCI combustion. 
     With an increase in the load of the engine  1 , the amount of the fuel injection and the temperature in the combustion chamber  17  increase, which creates a condition easily starting the CI combustion earlier. That is, a higher load of the engine  1  tends to cause abnormal combustion, such as pre-ignition and knocking, of the air-fuel mixture. However, as described above, the temperature on the periphery of the combustion chamber  17  decreases due to the latent heat of fuel vaporization. This reduces start of the CI combustion immediately after the spark ignition of the air-fuel mixture. 
     At a high load of the engine  1  in the SPCCI combustion, the stratification of the air-fuel mixture inside the combustion chamber  17  and the generation of a strong swirl flow inside the combustion chamber  17  allow sufficient SI combustion before the start of the CI combustion. This results in reduction in combustion noise and an excessive increase in combustion temperature to reduce NOx. In addition, the variation in the torque among the cycles decreases. 
     A lower temperature on the periphery of the combustion chamber  17  makes the CI combustion more moderate, which is advantageous in reducing combustion noise. In addition, since the CI combustion requires a shorter combustion period, torque and thermal efficiency improve at a high load of the engine  1 . Therefore, in the engine  1 , the SPCCI combustion is performed within a high load range, thereby improving fuel efficiency while reducing combustion noise. 
     &lt;Engine Control in CI Range ( 2 )&gt; 
     At a high speed of the engine  1 , a shorter time is required for the crank angle to change by 1°. For example, in the high speed and high load range, it is thus difficult to stratify the air-fuel mixture in the combustion chamber  17  to perform the SPCCI combustion. On the other hand, since the engine  1  has a high geometric compression ratio. Thus, if the SI combustion is to be performed particularly in a high load range, abnormal combustion such as knocking may occur. To address the problem, the engine  1  performs a new CI combustion mode in a higher speed operation in the CI range ( 2 ). The CI range ( 2 ) extends over all the ranges in the load direction from low to high loads. 
     This CI combustion utilizes what is called a “broken reaction zone”. In the broken reaction zone, the conditions inside the combustion chamber  17  are as follows. A lean air-fuel mixture and/or a strong flow in the combustion chamber  17  do(es) not allow the progress of the combustion by the flame propagation, even after the spark plug  25  has ignited the air-fuel mixture. The combustion mode in the CI range ( 2 ) is based on the following new finding obtained when the air-fuel mixture was viewed microscopically. If the spark plug  25  ignites the air-fuel mixture in the broken reaction zone, the flame does not go out but is stored while being unable to cause the flame propagation. 
     In  FIG. 6 , reference character  604  denotes an example including fuel injection times (reference characters  6041  and  6043 ), an ignition time (reference character  6042 ), and a combustion waveform (reference character  6044 ), in a high load operating state  604  of the engine  1  in the SPCCI range ( 2 ). 
     The air-fuel ratio (A/F) of the air-fuel mixture is basically equal to the stoichiometric air-fuel ratio (i.e., A/F=14.7) in the entire combustion chamber  17  in an operation of the engine  1  in the CI range ( 2 ). The excessive air ratio λ of the air-fuel mixture may be 1.0±0.2. In a high load range including the all loads within the CI range ( 2 ), the excessive air ratio λ of the air-fuel mixture may be lower than 1. 
     As shown in the map  501  in  FIG. 5 , in an operation of the engine  1  in the CI range ( 2 ), the supercharger  44  is turned on (see S/C ON) throughout all the ranges of the engine  1  to increase the supercharging pressure. 
       FIG. 9  illustrates a change in the flow strength in the combustion chamber  17  from the intake stroke to the compression stroke. In an operation of the engine  1  in the CI range ( 2 ), the swirl control valve  56  is fully open as shown in the map  502  in  FIG. 5 . Accordingly, no swirl flow but only a tumble flow occurs in the combustion chamber  17 . Such full opening of the swirl control valve  56  improves the filling efficiency, while reducing pump losses at a high speed of the engine  1 . 
     As the intake air flows into the combustion chamber  17  in the intake stroke, a tumble flow occurs and the flow in the combustion chamber  17  gradually strengthens. The flow in the combustion chamber  17  that has strengthened in the intake stroke weakens once at the post-stage of the intake stroke. However, with a lift of the piston  3  toward top dead center in the compression stroke, the flow in the combustion chamber  17  strengthens again due to what is called a “spin-up phenomenon”. As indicated by the white arrow in  FIG. 9 , the flow in the combustion chamber  17  has a predetermined strength (see the broken line) or more in the specific period. The specific period extends from the start of a middle stage, where the compression stroke is divided into three stages of an initial stage, a middle stage, and a final stage, to the end of a post-mid stage, where the compression stroke is divided into four stages of a pre-stage, a pre-mid stage, a post-mid stage, and a post-stage. In the specific period, the tumble ratio in the combustion chamber  17  is equal to or higher than a predetermined value. The “tumble ratio” is the value obtained by dividing the angular velocity ω of the intake air around an axis by the angular velocity ωc of the crankshaft  15 . The axis is parallel to the crankshaft  15  passing through the center of gravity of the combustion chamber  17 , whose position varies depending on a change in the volume of the combustion chamber  17 . The angular velocity ω of the intake air can be obtained as follows. Specifically, the inside of the combustion chamber  17  is divided into a large number of tiny regions corresponding to tiny crank angles from the start of the intake stroke to the end of the compression stroke. The angular momentum L of the mass points (air) of the tiny regions around the axis and the inertia momenta I of the mass points of the tiny regions are obtained. The angular momenta L of all the tiny regions are summed up throughout the tiny crank angles. The sum is divided by the sum of the inertia momenta I of all the tiny regions throughout the tiny crank angles. As a result, the angular velocity ω of the intake air can be obtained. 
     Formation of the air-fuel mixture with a low fuel concentration with respect to the stoichiometric air-fuel ratio in the combustion chamber  17  in or before the specific period does not allow combustion by the flame propagation, even after the spark plug  25  ignites the air-fuel mixture in the specific period. That is, in the CI range ( 2 ), the spark plug  25  ignites the air-fuel mixture when the conditions inside the combustion chamber  17  fall within the broken reaction zone. Accordingly, the flame is stored in the combustion chamber  17  without starting the combustion by the flame propagation. 
     After that, as the crank angle advances, the flow in the combustion chamber  17  weakens. In addition, the fuel is additionally supplied into the combustion chamber  17  to increase the fuel concentration of the air-fuel mixture. Then, the conditions inside the combustion chamber  17  fall out of the broken reaction zone. At the post-stage of the compression stroke, motoring increases the temperature and pressure inside the combustion chamber  17 . As a result, the flame stored starts the combustion of the air-fuel mixture at the post-stage of the compression stroke or in the expansion stroke. 
     Next, fuel injection control and ignition control will be described in detail with reference to  FIG. 6 . In an operation of the engine  1  in the CI range ( 2 ), the injector  6  performs the fuel injection (i.e., the first fuel injection  6041 ) in the intake stroke. The first fuel injection may be performed at once or in a divided manner, for example. The start of the fuel injection in the intake stroke allows for formation of a homogeneous or substantially homogeneous air-fuel mixture in the combustion chamber  17 . The air-fuel mass ratio A/F or a gas-fuel mass ratio G/F, in which the gas includes air, of the air-fuel mixture formed at this time is higher than the stoichiometric air-fuel ratio. The amount of the first fuel injection  6041  is determined by the load of the engine  1  and the division ratio between the first fuel injection  6041  and the second fuel injection  6043  described later. Note that the fuel injection period varies depending on the amount of the first fuel injection  6041 . The start of the first fuel injection  6041  may be set as appropriate in accordance with the amount of the first fuel injection  6041  so as to form the air-fuel mixture with a low fuel concentration with respect to the stoichiometric air-fuel ratio in the combustion chamber  17  at least in or before the specific period described above. 
     After the end of the first fuel injection, the spark plug  25  ignites the air-fuel mixture (see the reference character  6042 ). An upper illustration  1001  of  FIG. 10  illustrates an ignition time in a high load operation of the engine  1  in the CI range ( 2 ). The speed N 1  in  FIG. 10  corresponds to the speed N 1  of the map shown in  FIG. 5 . In the upper illustration  1001  of  FIG. 10 , the vertical axis represents the crank angle, which advances toward the top of the vertical axis. 
     In the upper illustration  1001  of  FIG. 10 , the hatched range indicates the ignition time of the spark plug  25 . The spark plug  25  ignites the air-fuel mixture at an appropriate timing within the hatched range. 
     The spark plug  25  ignites the air-fuel mixture, within the compression stroke, at or before the post-mid stage. This allows for storage of the flame in the combustion chamber  17  without allowing the progress of the combustion by the flame propagation. The flame is dispersed or diffused by the flow in the combustion chamber  17 . As is apparent from  FIG. 9 , the ignition at a too early or late phase within the compression stroke means that the air-fuel mixture is ignited when the flow inside the combustion chamber  17  is weak. Thus, the ignition time has an advance limit (i.e., the upper line in the upper illustration  1001 ) and a retard limit (i.e., the lower line in the upper illustration  1001 ) so that the spark plug  25  ignites the air-fuel mixture, when the conditions inside the combustion chamber  17  fall within the broken reaction zone, as shown in the upper illustration  1001  of  FIG. 10 . The spark plug  25  may ignite the air-fuel mixture at the middle stage, for example, where the compression stroke is divided into three stages of an initial stage, a middle stage, and a final stage. 
     The ignition of the air-fuel mixture by the spark plug  25  in the CI range ( 2 ) is much more advanced than minimum advance for best torque (MBT), which can be set where the SI combustion is performed in the same operating state as illustrated by the one-dot-chain lines in the upper illustration  1001  of  FIG. 10 . The MBT is found at the post-stage of the compression stroke, for example. 
     The ignition time may vary in accordance with to the magnitude of the speed of the engine  1 . In the example of the upper illustration  1001 , the ignition time advances with an increase in the speed of the engine  1 . The variation in the advance limit with an increase in the speed of the engine  1  is larger than the variation in the retard limit. That is, the upper line has a steeper slope than the lower line in the upper illustration  1001 . 
     The spark plug  25  may perform a plurality of ignitions within the specific period. This increases the number of flames generated in the combustion chamber  17  and allows a strong flow in the combustion chamber  17  to diffuse the large number of flames into the combustion chamber  17 . Accordingly, the ignitability of the air-fuel mixture improves and the combustion period of the air-fuel mixture further shortens. 
     In the period of the compression stroke after the spark plug  25  has ignited the air-fuel mixture, the injector  6  injects the fuel into the combustion chamber  17  (i.e., the second fuel injection  6043 ). The fuel concentration of the air-fuel mixture in the combustion chamber  17  increases. The second fuel injection  6043  makes the A/F or G/F of the air-fuel mixture lower than or equal to the stoichiometric air-fuel ratio. With an increase in the fuel concentration of the air-fuel mixture, the flow in the combustion chamber  17  weakens, whereby the conditions inside the combustion chamber  17  fall out of the broken reaction zone. In addition, the temperature and the pressure inside the combustion chamber  17  are increased by motoring as the piston comes closer to compression top dead center. Accordingly, near compression top dead center, the flame stored starts the combustion of the air-fuel mixture by the autoignition at once (see the reference character  6044 ). The center of gravity of this combustion is closer to compression top dead center, which improves the thermal efficiency of the engine  1 . In addition, this combustion mode requires a shorter combustion period and thus reduces knocking. 
     In accordance with the fuel concentration of the air-fuel mixture in the combustion chamber  17 , whether or not the conditions inside the combustion chamber  17  fall out of the broken reaction zone changes. A change in the time when the conditions of the combustion chamber  17  fall out of the broken reaction zone changes the start of the combustion of the air-fuel mixture. The amount of the second fuel injection  6043  may thus be adjusted as appropriate to start the combustion of the air-fuel mixture at an appropriate timing. 
     As illustrated in  FIG. 6 , the second fuel injection  6043  may be performed after the spark plug  25  has ignited the air-fuel mixture. A delay in the end of the fuel injection  6043  shortens the vaporization time of the fuel injected in the second fuel injection  6043 , which may be disadvantageous in terms of exhaust gas emission performance or fuel efficiency. The start of the second fuel injection  6043  may be set as appropriate based on the amount of the second fuel injection  6043  not to cause a delay of the injection end of the second fuel injection  6043 . For example, the lower illustration  1002  of  FIG. 10  illustrates the start of the second fuel injection  6043 . In the lower illustration  1002  of  FIG. 10 , the vertical axis of represents the crank angle, which advances toward the top of the vertical axis. In the lower illustration  1002  of the figure, the hatched area represents the start of the second fuel injection  6043 . The injector  6  performs the second fuel injection  6043  at an appropriate timing within the hatched range. 
     With a change in the start of the second fuel injection  6043 , the time when the conditions inside the combustion chamber  17  fall out of the broken reaction zone changes. Thus, with a change in the start of the second fuel injection  6043  by the injector  6 , the start of the combustion of the air-fuel mixture changes. The start of the second fuel injection  6043  may be adjusted to start the combustion of the air-fuel mixture at an appropriate timing. 
     The start of the second fuel injection  6043  may change in accordance with the magnitude of the speed of the engine  1 . Specifically, with an increase in the speed of the engine  1 , the ignition timing advances. In accordance with the advance, the start of the second fuel injection  6043  may advance. In the example of the lower illustration  1002 , the start of the second injection advances with an increase in the speed of the engine  1 . The variation in the advance limit (i.e., the slope of the upper line in the lower illustration  1002 ) with an increase in the speed of the engine  1  is larger than the variation in the retard limit (i.e., the slope of the lower line in the lower illustration  1002 ). 
     The ratio between the amounts of the first and second injection  6041  and  6043  may be set as appropriate. 
     The EGR system  55  introduces the external and/or internal EGR gas into the combustion chamber  17  in an operation of the engine  1  in the CI range ( 2 ). The EGR rate in the combustion chamber  17  may be adjusted to start the combustion of the air-fuel mixture at an appropriate timing. 
     In the CI range ( 2 ), water may be injected into the combustion chamber  17  at an appropriate timing in the compression stroke to start the combustion of the air-fuel mixture at an appropriate timing. 
     In the maps  501  and  502  shown in  FIG. 5 , the combustion mode described above may be employed throughout the entire CI range ( 2 ) or the load range on the R-L line in the CI area ( 2 ). 
     &lt;Control Process of Engine&gt; 
       FIG. 11  illustrates a flowchart of the control of the engine  1  described above and executed by the ECU  10 . 
     First, in step S 1  after the start of the flow, the ECU  10  reads signals of the various sensors SW 1  to SW 16 . In the subsequent step S 2 , the ECU  10  determines whether or not the operating state of the engine  1  is within the high speed range. The high speed range corresponds to the CI range ( 2 ) described above. In step S 2 , the ECU  10  may determine whether or not the engine speed is equal to or higher than N 1 . If the determination in step S 2  is YES, the control process proceeds to step S 3 . If the determination in the step S 2  is NO, the process proceeds to a step S 6 . 
     Steps S 3  to S 5  are the control steps within the CI range ( 2 ). The steps S 3  to S 5  proceed in this order. 
     First, in step S 3 , the ECU  10  executes the first fuel injection by the injector  6 . Accordingly, leaner air-fuel mixture with respect to the stoichiometric air-fuel ratio is generated in the combustion chamber  17 . In the subsequent step S 4 , the ECU  10  executes ignition by the spark plug  25  at a predetermined time. As described above, the spark plug  25  performs the ignition before the post-mid stage of the compression stroke. In step S 5 , the ECU  10  executes then the second fuel injection by the injector  6 . As a result, near compression top dead center, the flame made in advance in the combustion chamber  17  starts the combustion of the air-fuel mixture at once by autoignition. 
     On the other hand, in step S 6 , the control in the SPCCI range ( 1 ) is performed in accordance with the load of the engine  1 . 
     &lt;Variations of Engine&gt; 
     Note that application of the disclosed technique is not limited to the engine  1  with the configuration described above. The engine may employ various configurations. 
       FIG. 12  shows a configuration of an engine  100  according to a variation. The engine  100  includes a turbocharger  70  in place of the mechanical supercharger  44 . 
     The turbocharger  70  includes a compressor  71  in the intake passage  40  and a turbine  72  in the exhaust passage  50 . The turbine  72  is rotated by the exhaust gas flowing through the exhaust passage  50 . The compressor  71  is rotated by the rotational drive of the turbine  72  and supercharges the gas in the intake passage  40  to be introduced into the combustion chamber  17 . 
     The exhaust passage  50  includes an exhaust bypass passage  73 . The exhaust bypass passage  73  connects parts of the exhaust passage  50  upstream and downstream of the turbine  72  as to bypass the turbine  72 . The exhaust bypass passage  73  includes a waste gate valve  74 . The waste gate valve  74  adjusts the flow rate of the exhaust gas flowing through the exhaust bypass passage  73 . 
     In this exemplary configuration, the turbocharger  70 , the bypass passage  47 , the air bypass valve  48 , the exhaust bypass passage  73 , and the waste gate valve  74  constitute a supercharging system  49  in the intake and exhaust passages  40  and  50 . 
     The engine  100  switches opening/closing of the air bypass valve  48  and the waste gate valve  74  to or not to cause the turbocharger  70  to supercharge the gas to be introduced into the combustion chamber  17 . 
     When the gas introduced into the combustion chamber  17  is not supercharged, the waste gate valve  74  opens. Accordingly, the gas flowing through the exhaust passage  50  bypasses the turbine  72 , that is, passes not through the turbine  72  but through the bypass passage  73  into the catalyst converters. Then, the turbine  72  does not receive the flow of the exhaust gas and thus does not drive the turbocharger  70 . At this time, the air bypass valve  48  is fully open. As a result, the gas flowing through the intake passage  40  flows through none of the compressor  71  or the intercooler  46  but through the bypass passage  47  into the surge tank  42 . 
     When the gas to be introduced into the combustion chamber  17  is supercharged, the waste gate valve  74  is not fully open but closed slightly. Accordingly, at least a portion of the exhaust gas flowing through the exhaust passage  50  passes through the turbine  72  and flows to the catalyst converters. Then, the turbine  72  rotates upon receipt of the exhaust gas and drives the turbo turbocharger  70 . When the turbocharger  70  is driven, the gas in the intake passage  40  is supercharged by the rotation of the compressor  71 . At this time, if the air bypass valve  48  is open, a portion of the gas that has passed through the compressor  71  flows back from the surge tank  42  through the bypass passage  47  to the upstream side of the compressor  71 . The supercharging pressure of the gas inside the intake passage  40  can be controlled by adjusting the opening degree of the air bypass valve  48  as in the case using the mechanical supercharger described above. 
     Whether or not the turbocharger  70  supercharges the gas in the intake passage  40  may be determined in accordance with a map  503  shown in  FIG. 13 , for example. That is, the turbocharger  70  may not perform the supercharging in the low load range in the SPCCI range ( 1 )(see T/C OFF), whereas the turbocharger  70  may perform the supercharging in the medium and high load ranges in the SPCCI range ( 1 ) as well as in the CI range ( 2 ) (see T/C ON). In the low load range, a lower torque is required. The supercharging is thus less needed and the air-fuel mixture is leaner with respect to the stoichiometric air-fuel ratio. As a result, the temperature of the exhaust gas decreases. In order to maintain the three-way catalysts  511  and  513  at the activation temperatures, the waste gate valve  74  may open to bypass the turbine  72 , thereby reducing heat dissipation in the turbine  72  and supplying hot exhaust gas to the three-way catalysts  511  and  513 . 
     The operation of the engine  100  including the turbocharger  70  may be controlled in accordance with the flowchart shown in  FIG. 11 . The engine  100  also can improve thermal efficiency, while reducing abnormal combustion at a high speed of the engine  100 . 
     Although not shown, the disclosed technology is applicable to a naturally aspirated engine without any supercharger. 
     &lt;Other Exemplary Configuration&gt; 
     With the above-described configuration, the second fuel injection  6043  is performed after the ignition of the air-fuel mixture in the operation of engine  1  in the CI range ( 2 ). However, the present disclosure is not limited thereto. The second fuel injection  6043  may be omitted. The broken reaction zone relates to two parameters of the air-fuel ratio of the air-fuel mixture and the flow strength in the combustion chamber  17 . However, there may be a case where the conditions inside the combustion chamber  17  fall out of the broken reaction zone merely when the flow strength in the combustion chamber  17  weakens as the piston  3  comes closer to compression top dead center. As a result, near the compression top dead center, autoignition of the flame occurs to start the combustion of the mixture even when the second fuel injection  6043  is not performed. 
     In the configuration described above, the first fuel injection  6041  generates the homogeneous or substantially homogeneous air-fuel mixture with the A/F or G/F higher than the stoichiometric air-fuel ratio in the combustion chamber  17 . The time of the first fuel injection  6041  may be adjusted to locally generate the air-fuel mixture with an A/F or a G/F higher than the stoichiometric air-fuel ratio near the spark plug  25  at the ignition timing. 
     In addition to the injector for directly injecting the fuel into the combustion chamber  17 , a port injector facing the inside of the intake port may be provided. In particular, the first fuel injection for injecting the fuel in the intake stroke may be performed by the port injector. 
     The engine  1  may include an ignition unit that causes an arc discharge or a plasma discharge in each combustion chamber  17  in place of the spark plug  25  that performs the spark discharge. 
     DESCRIPTION OF REFERENCE CHARACTERS 
     
         
         
           
               1 ,  100  Engine 
               6  Injector (Fuel Supply Unit) 
               10  ECU (Control Unit) 
               17  Combustion Chamber 
               25  Spark Plug (Ignition Unit) 
             S 3  First Fuel Supply Step 
             S 4  Ignition Step 
             S 5  Second Fuel Supply Step