Patent Description:
Patent Document <NUM> discloses an engine that burns air-fuel mixture in a combustion chamber by compression ignition in a predetermined low load and low speed range. The engine burns 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 air-fuel mixture by spark ignition using a spark plug near a compression top dead center also in the predetermined region.

Patent Document <NUM> discloses an engine that burns 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 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 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. Patent Document <NUM> discloses an injector that injects fuel into a combustion chamber, a spark plug that emits a spark from an electrode exposed to the combustion chamber, and a control unit that controls the operation of the injector and the spark plug. Patent Document <NUM> also relates to a spark-ignition petrol engine in which compression ignition (CI) combustion is performed under a predetermined condition in which a mixture of fuel and air injected from the injector is combusted by auto-ignition. Thereby, proper CI combustion is realised without increasing combustion noise and soot, while self-ignition of the air-fuel mixture is promoted by ignition assistance. Further disclosed is a spark ignition gasoline engine capable of quickly creating a temperature condition suitable for CI combustion.

In an engine with a geometric compression ratio increased for the main purpose of improving the thermal efficiency, an ignition unit performs spark ignition of 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 the compression top dead center. Accordingly, the thermal efficiency of the engine decreases.

The present disclosure increases the thermal efficiency of an engine.

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 do(es) 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 air-fuel mixture is to be burned by the flame propagation, made flame goes out. In the typical engine control, 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 flame stored starts to cause combustion of the air-fuel mixture at once. Based on the finding that the flame stored 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 as defined in claim <NUM> of the invention 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 step of creating, using a tumble port, a strong tumble flow in the combustion chamber, a first fuel supply step and an ignition step. In the first fuel supply step, the fuel supply unit supplies the fuel into the combustion chamber so as to generate an air-fuel mixture having an air-fuel mass ratio higher than a stoichiometric air-fuel ratio. In the ignition step, an ignition unit arranged in the combustion chamber makes flame after the supply of the fuel into the combustion chamber in the first fuel supply step and at a timing when the tumble ratio in the combustion chamber is equal to or higher than a predetermined value in a compression stroke during or before a post-mid stage such that the lean air-fuel mixture and the strong flow in the combustion chamber do not allow the progress of the combustion by flame propagation and the flame does not go out but is stored 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 a sequential manner. After the ignition step, a second fuel supply step in which the fuel supply unit supplies fuel into the combustion chamber to increase a fuel concentration of the air-fuel mixture in the combustion chamber, and a step in which combustion of the air-fuel mixture is started by the stored flame by auto-ignition after the second fuel supply step.

According to the above configuration, in the first fuel supply step, fuel is supplied into the combustion chamber to generate air-fuel mixture in the combustion chamber. In the first 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 first fuel supply step, the ignition step is executed. In the ignition step, the ignition unit makes 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 a 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 the 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 the piston moves toward the 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, 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 flame is made, it is possible to keep the flame as it 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 the 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 increases 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 the compression top dead center. The center of gravity of this combustion is close to the 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 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 in a compression stroke.

The intake port is configured as a so-called tumble port. Accordingly, when tumble flow (i.e., vertical vertex) 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 spin-up phenomenon in a period from the initial stage to the middle stage of the compression stroke. The ignition unit makes 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 makes 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 fuel in the combustion chamber at a timing when the ignition unit makes the flame in the combustion chamber so that 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 progress of combustion by flame propagation. If 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 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 lean air-fuel mixture with low fuel concentration, thereby making it possible to reduce the risk of preignition.

In the engine control method according to the invention, after the ignition step, a second fuel supply step in which the fuel supply unit supplies fuel into the combustion chamber to increase a fuel concentration of air-fuel mixture in the combustion chamber.

In the second fuel supply step, additional fuel supply by the fuel supply unit leads to 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 <NUM> or more. The engine control method disclosed herein improves the thermal efficiency, while reducing abnormal combustion in an engine with a high compression ratio.

The present invention also relates to an engine control device according to claim <NUM>. 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 flame after the supply of the fuel into the combustion chamber by the fuel supply unit and at a timing when a tumble flow in the combustion chamber is equal to or higher than a predetermined value in a compression stroke during or before a post-mid stage such that the lean air-fuel mixture and the strong flow in the combustion chamber do not allow the progress of the combustion by flame propagation and the flame does not go out but is stored 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 fuel supply unit is configured to supply, after the ignition unit has made the flame, fuel into the combustion chamber to increase a fuel concentration of air-fuel mixture in the combustion chamber.

The engine may have a geometric compression ratio of <NUM> or more.

The engine control method and device described above increase the thermal efficiency of an engine.

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> illustrates a configuration of an engine system including an engine <NUM>. <FIG> illustrates a configuration of a combustion chamber <NUM>. In this <FIG>, the upper illustration corresponds to a top view of the combustion chamber <NUM>, whereas the lower illustration is a cross-sectional view taken along the line II-II of the upper illustration. <FIG> illustrates configurations of the combustion chamber <NUM> and an intake system. <FIG> is a block diagram illustrating a configuration of an engine control device. In <FIG>, 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 <FIG> and <FIG>, 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 <NUM> 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 <NUM>. The engine <NUM> is mounted in a four-wheeled motor vehicle. The motor vehicle travels in accordance with the operation of the engine <NUM>. Fuel of the engine <NUM> is gasoline in this exemplary configuration. The fuel may be gasoline containing bioethanol, for example. The fuel for the engine <NUM> may be any fuel as long as the fuel is liquid fuel which contains at least gasoline.

The engine <NUM> is a multi-cylinder engine. As shown in <FIG>, the engine <NUM> includes an engine body <NUM> with the combustion chamber <NUM>. The engine body <NUM> includes a cylinder block <NUM> and a cylinder head <NUM> above the cylinder block <NUM>. Inside the cylinder block <NUM>, a plurality of cylinders <NUM> are arranged. Note that <FIG> and <FIG> each show only one of the cylinders <NUM>.

A piston <NUM> is slidably fitted into each of the cylinders <NUM>. The piston <NUM> is coupled to a crankshaft <NUM> via a connecting rod <NUM>. The piston <NUM> defines the combustion chamber <NUM>, together with each of the cylinders <NUM> and the cylinder head <NUM>. Note that the "combustion chamber" meant here is not limited to the space defined when the piston <NUM> 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 <NUM>, the cylinder <NUM>, and the cylinder head <NUM>, regardless of the position of the piston <NUM>. 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>, the lower surface of the cylinder head <NUM>, that is, the ceiling of the combustion chamber <NUM>, includes inclined surfaces <NUM> and <NUM>. The inclined surface <NUM> is an upward slope extending from the intake side toward an injection axis X2 of an injector <NUM>, which will be described later. On the other hand, the inclined surface <NUM> is an upward slope extending from the exhaust side toward the injection axis X2. The ceiling of the combustion chamber <NUM> is in the shape of what is called a "pent roof".

The upper surface of the piston <NUM> protrudes toward the ceiling of the combustion chamber <NUM>. Defined above the upper surface of the piston <NUM> is a cavity <NUM>. The cavity <NUM> is recessed from the upper surface of the piston <NUM>. The cavity <NUM> faces the injector <NUM> which will be described later. The center of the cavity <NUM> is shifted toward the exhaust side with respect to a center axis X1 of the cylinder <NUM> but agrees with the injection axis X2 of the injector <NUM>.

The cavity <NUM> includes a projection <NUM>. The projection <NUM> is located on the injection axis X2 of the injector <NUM>. This projection <NUM> is in the shape of a substantial cone extending upward from the bottom of the cavity <NUM> toward the ceiling of the combustion chamber <NUM>. The cavity <NUM> is symmetric about the injection axis X2 of the injector <NUM>.

The cavity <NUM> also includes a recess <NUM> around the projection <NUM>. The recess <NUM> surrounds the entire circumference of the projection <NUM>. The circumferential side surface of the recess <NUM> is inclined with respect to the injection axis X2 from the bottom of the cavity <NUM> toward the opening of the cavity <NUM>. The inner diameter of the cavity <NUM> of the recess <NUM> gradually increases from the bottom of the cavity <NUM> toward the opening of the cavity <NUM>.

Note that the shape of the combustion chamber <NUM> is not limited to that illustrated in <FIG>. That is, the shapes of the cavity <NUM>, the upper surface of the piston <NUM>, and the ceiling of the combustion chamber <NUM>, for example, may be changed as appropriate. For example, the cavity <NUM> may be symmetric about the center axis X1 of the cylinder <NUM>. The inclined surfaces <NUM> and <NUM> may be symmetric about the center axis X1 of the cylinder <NUM>. The cavity <NUM> may have a shallow bottom that is shallower than the bottom of the recess <NUM> so as to face a spark plug <NUM>, which will be described later.

The engine <NUM> has a geometric compression ratio ranging from <NUM> to <NUM>. As will be described later, the engine <NUM> 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 air-fuel mixture in the combustion chamber <NUM>. The CI combustion is started by autoignition of the air-fuel mixture in the combustion chamber <NUM>. The combined mode of these SI and CI combustions is as follows. The air-fuel mixture in the combustion chamber <NUM> 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 <NUM>. In this engine <NUM>, there is no need to significantly increase the temperature of the combustion chamber <NUM> when the piston <NUM> reaches the compression top dead center, that is, the compression end temperature, for the autoignition of the air-fuel mixture.

The cylinder head <NUM> includes two intake ports <NUM> for each cylinder <NUM>. As shown in <FIG>, the intake ports <NUM> include two intake ports of first and second intake ports <NUM> and <NUM>. The first and second intake ports <NUM> and <NUM> are aligned along the axis of the crankshaft <NUM>, that is, in the front-rear direction of the engine body <NUM>. The intake ports <NUM> communicate with the combustion chamber <NUM>. Although not shown in detail, the intake ports <NUM> are what are called "tumble ports". That is, each intake port <NUM> is in a shape causing a tumble flow in the combustion chamber <NUM> in the intake stroke.

Each intake port <NUM> includes an intake valve <NUM>. The intake valve <NUM> opens and closes the intake port <NUM> between the combustion chamber <NUM> and the intake port <NUM>. The engine <NUM> includes a valve train mechanism for the intake valves <NUM>. The intake valves <NUM> are opened and closed by a valve train mechanism at predetermined timing. The valve train mechanism for the intake valves <NUM> 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>, an intake electric sequential-valve timing (S-VT) <NUM>. The intake electric S-VT <NUM> is a variable valve train mechanism of a phase type that causes the intake valves <NUM> to open at a constant angle and to open and close at variable times. The intake electric S-VT <NUM> continuously changes the rotational phase of an exhaust camshaft within a predetermined angular range. Accordingly, the opening and closing times of the intake valve <NUM> change continuously. Note that the valve train mechanism for the intake valves <NUM> may include a hydraulic S-VT instead of the electric S-VT. The valve train mechanism for the intake valves <NUM> may include a variable valve train mechanism that changes the amounts of lift of the intake valves <NUM> and/or a variable valve train mechanism that changes the opening angles (or the opening periods) of the intake valves <NUM>.

The cylinder head <NUM> also includes two exhaust ports <NUM> for each cylinder <NUM>. As shown in <FIG>, the exhaust ports <NUM> include two exhaust ports, i.e., a first exhaust port <NUM> and a second exhaust port <NUM>. The first and second exhaust ports <NUM> and <NUM> are aligned in the front-rear direction of the engine body <NUM>. The exhaust ports <NUM> communicate with the combustion chamber <NUM>.

Each exhaust port <NUM> includes an exhaust valve <NUM>. The exhaust valve <NUM> opens and closes the exhaust port <NUM> between the combustion chamber <NUM> and the exhaust port <NUM>. The engine <NUM> includes a valve train mechanism for the exhaust valves <NUM>. The exhaust valves <NUM> are opened and closed by the valve train mechanism at predetermined timing. The valve train mechanism for the exhaust valves <NUM> 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>, an exhaust electric S-VT <NUM>. The exhaust electric S-VT <NUM> is a variable valve train mechanism of a phase type that causes the exhaust valves <NUM> to open at a constant angle and to open and close at variable times. The exhaust electric S-VT <NUM> continuously changes the rotational phase of an exhaust camshaft within a predetermined angular range. Accordingly, the opening and closing times of the exhaust valves <NUM> change continuously. Note that the valve train mechanism for the exhaust valves <NUM> may include a hydraulic S-VT instead of the electric S-VT. The valve train mechanism for the exhaust valves <NUM> may include a variable valve train that changes the amounts of lift of the exhaust valves <NUM> and/or a variable valve train mechanism that changes the opening angles (or the opening periods) of the exhaust valves <NUM>.

In the engine <NUM>, the intake and exhaust electric S-VTs <NUM> and <NUM> adjust the length of the overlap period between the opening times of the intake valves <NUM> and the closing times of the exhaust valves <NUM>. Accordingly, hot burned gas is confined in the combustion chamber <NUM>. That is, internal exhaust gas recirculation (EGR) gas is introduced into the combustion chamber <NUM>. The adjustment of the length of the overlap period allows scavenge of the residual gas (burned gas) in the combustion chamber <NUM>.

The cylinder head <NUM> includes the injector <NUM> for each cylinder <NUM>. The injector <NUM> directly injects the fuel into the combustion chamber <NUM>. The injector <NUM> is an example fuel supply unit. The injector <NUM> faces the inside of the combustion chamber <NUM> at the valley of the pent roof, at which the inclined surface <NUM> on the intake side and the inclined surface <NUM> on the exhaust side intersect each other, and is opposed to the cavity <NUM>.

As shown in <FIG>, the injection axis X2 of the injector <NUM> is parallel to the center axis X1 of the cylinder <NUM> and is closer to the exhaust side than the center axis X1 of the cylinder <NUM>. The injection axis X2 of the injector <NUM> agrees with the position of the projection <NUM> of the cavity <NUM>. Note that the injection axis X2 of the injector <NUM> may agree with the center axis X1 of the cylinder <NUM>. In this case as well, it is desirable that the injection axis X2 of the injector <NUM> agree with the position of the projection <NUM> of the cavity <NUM>.

Although not shown in detail, the injector <NUM> includes a multi-port combustion injection valve with a plurality of nozzle ports. As indicated by the two-dot chain lines in <FIG>, the injector <NUM> injects the fuel so that the fuel spray spreads radially from the center of the combustion chamber <NUM> and spreads obliquely downward from the ceiling of the combustion chamber <NUM>.

In this exemplary configuration, the injector <NUM> has ten nozzle ports. The nozzle ports are arranged at equal angles along the circumference of the injector <NUM>. As shown in the upper illustration of <FIG>, the axes of the nozzle ports are shifted along the circumference of the injector <NUM> with respect to the spark plug <NUM> which will be described later. That is, the spark plug <NUM> 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 <NUM> with the spark plug <NUM>, thereby making it possible to avoid making any electrode wet.

A fuel supply system <NUM> is connected to the injector <NUM>. The fuel supply system <NUM> includes a fuel tank <NUM> configured to store the fuel, and a fuel supply passage <NUM> that connects the fuel tank <NUM> and the injector <NUM> together. The fuel supply passage <NUM> includes a fuel pump and a common rail <NUM>. The fuel pump <NUM> pumps out the fuel to the common rail <NUM>.

In this exemplary configuration, the fuel pump <NUM> is a plunger pump driven by the crankshaft <NUM>. The common rail <NUM> stores the fuel pumped out from the fuel pump <NUM> at a high fuel pressure. When the injector <NUM> opens, the fuel stored in the common rail <NUM> is injected from a nozzle port of the injector <NUM> into the combustion chamber <NUM>.

The fuel supply system <NUM> can supply fuel to the injector <NUM> at a high pressure of <NUM> MPa or more. The maximum fuel pressure of the fuel supply system <NUM> may be about <NUM> MPa, for example. The pressure of the fuel to be supplied to the injector <NUM> may vary in accordance with the operating state of the engine <NUM>. Note that the configuration of the fuel supply system <NUM> is not limited to the configuration described above.

The cylinder head <NUM> includes the spark plug <NUM> attached to each of the cylinders <NUM>. The spark plug <NUM> performs a spark discharge between electrodes arranged in the combustion chamber <NUM>, thereby forcibly igniting the air-fuel mixture in the combustion chamber <NUM>. The spark plug <NUM> is an example of an ignition unit.

In this exemplary configuration, as shown in <FIG>, the spark plug <NUM> is closer to the intake side with respect to the center axis X1 of the cylinder <NUM> in the combustion chamber <NUM>. The spark plug <NUM> is adjacent to the injector <NUM> and interposed between the two intake ports. In addition, the spark plug <NUM> is attached to the cylinder head <NUM>, while being inclined such that the bottom of the plug is closer to the center of the combustion chamber <NUM> than the top of the plug. The electrodes of the spark plug <NUM> face the combustion chamber <NUM> and are located near the ceiling of the combustion chamber <NUM>.

One side surface of the engine body <NUM> is connected to an intake passage <NUM>. The intake passage <NUM> communicates with the intake ports <NUM> of each cylinder <NUM> and with the combustion chamber <NUM> via the intake ports <NUM>. Through the intake passage <NUM>, the gas introduced into the combustion chamber <NUM> flows. Located at the upstream end of the intake passage <NUM> is an air cleaner <NUM> that filters fresh air. Located near the downstream end of the intake passage <NUM> is a surge tank <NUM>. A part of the intake passage <NUM> downstream of the surge tank <NUM> forms independent passages that branch off for the respective cylinders <NUM>. The downstream end of each independent passage is connected to the intake ports <NUM> of the associated one of the cylinders <NUM>.

A throttle valve <NUM> is interposed between the air cleaner <NUM> and the surge tank <NUM> in the intake passage <NUM>. The opening degree of the throttle valve <NUM> is adjusted to the amount of fresh air to be introduced into the combustion chamber <NUM>.

In the intake passage <NUM>, a supercharger <NUM> is provided downstream of the throttle valve <NUM>. The supercharger <NUM> supercharges the gas inside the intake passage <NUM> to be introduced into the combustion chamber <NUM>.

In the exemplary configuration, the supercharger <NUM> is a mechanical supercharger driven by the engine body <NUM>. The mechanical supercharger <NUM> may be of a Roots type, for example. The mechanical supercharger <NUM> may have any configuration. The mechanical supercharger <NUM> may be of a Lysholm type, a vane type, or a centrifugal type.

An electromagnetic clutch <NUM> is interposed between the supercharger <NUM> and the engine body <NUM>. The electromagnetic clutch <NUM> transmits a driving force from the engine body <NUM> to the supercharger <NUM> or blocks the driving force between the supercharger <NUM> and the engine body <NUM>. As will be described later, the supercharger <NUM> is turned on and off by an engine control unit (ECU) <NUM> that determines whether to engage or disengage the electromagnetic clutch <NUM>. Accordingly, the engine <NUM> determines whether or not to supercharge the gas to be introduced into the combustion chamber <NUM> by the supercharger <NUM>.

In the intake passage <NUM>, the intercooler <NUM> is provided downstream of the supercharger <NUM>. The intercooler <NUM> cools the gas compressed by the supercharger <NUM>. The intercooler <NUM> may be of a water-cooling type, for example. Alternatively, the intercooler <NUM> may be of an oil cooling type.

The intake passage <NUM> is also connected to a bypass passage <NUM>. The bypass passage <NUM> connects the upstream part of the supercharger <NUM> and the downstream part of the intercooler <NUM> in the intake passage <NUM> together so as to bypass the supercharger <NUM> and the intercooler <NUM>. The bypass passage <NUM> includes an air bypass valve <NUM>. The air bypass valve <NUM> adjusts the flow rate of gas flowing through the bypass passage <NUM>.

When the supercharger <NUM> is turned off, that is, when the electromagnetic clutch <NUM> is disengaged, the air bypass valve <NUM> fully opens. Accordingly, the gas flowing in the intake passage <NUM> bypasses the supercharger <NUM>, that is, passes through none of the supercharger <NUM> or the intercooler <NUM> but through the bypass passage <NUM> and flows into the surge tank <NUM>. The gas is then introduced into the combustion chamber <NUM> of the engine <NUM>. The engine <NUM> operates without supercharging, that is, with natural aspiration.

When the supercharger <NUM> is turned on, that is, when the electromagnetic clutch <NUM> is engaged, the gas flowing through the intake passage <NUM> passes through the supercharger <NUM> and the intercooler <NUM> and then flows into the surge tank <NUM>. At this time, if the air bypass valve <NUM> is open, a part of the gas that has passed through the supercharger <NUM> flows back from the surge tank <NUM> through the bypass passage <NUM> to the upstream side of the supercharger <NUM>. The back flow rate of such gas varies depending on the opening degree of the air bypass valve <NUM>. The supercharging pressure of the gas inside the intake passage <NUM> may be controlled by adjusting the opening degree of the air bypass valve <NUM>.

In this exemplary configuration, the supercharger <NUM>, the bypass passage <NUM>, and the air bypass valve <NUM> constitute a supercharging system <NUM> in the intake passage <NUM>.

The engine <NUM> includes a swirl generating unit that generates a swirl flow in the combustion chamber <NUM>. As shown in <FIG>, the swirl generating unit is a swirl control valve <NUM> attached to the intake passage <NUM>. The swirl control valve <NUM> is located in a secondary passage <NUM> out of primary and secondary passages <NUM> and <NUM> that communicate with the first and second intake ports <NUM> and <NUM>, respectively.

The swirl control valve <NUM> is an opening degree adjustment valve capable of narrowing the cross-section of the secondary passage <NUM>. In the combustion chamber <NUM>, a swirl flow occurs which has a strength corresponding to the opening degree of the swirl control valve <NUM>. The swirl flow circulates counterclockwise in <FIG> as indicated by the arrows (see the white arrows in <FIG> as well).

At a lower opening degree of the swirl control valve <NUM>, the flow rate of the intake air flowing from the first intake port <NUM> into the combustion chamber <NUM> relatively increases, whereas the flow rate of the intake air flowing from the second intake port <NUM> into the combustion chamber <NUM> relatively decreases, out of the first and second intake ports <NUM> and <NUM> aligned in the front-rear direction of the engine body <NUM>. This causes thus a stronger swirl flow in the combustion chamber <NUM>. When the swirl control valve <NUM> opens at a higher degree, the flow rates of the intake air flowing into the combustion chamber <NUM> from the first and second intake ports <NUM> and <NUM> are substantially equal to each other. This causes thus a weaker swirl flow in the combustion chamber <NUM>. When the swirl control valve <NUM> fully opens, the swirl flow does not occur.

Instead of or in addition to attaching the swirl control valve <NUM> to the intake passage <NUM>, the swirl generating section may employ the following configuration. The opening periods of two intake valves <NUM> to allow introduction of the intake air from only one of the intake valves <NUM> into the combustion chamber <NUM>. With the opening of only one of the two intake valves <NUM>, the intake air is unevenly introduced into the combustion chamber <NUM>, which allows generation of a swirl flow in the combustion chamber <NUM>. In addition, each intake port <NUM> may have an innovative shape so that the swirl generating unit generates a swirl flow in the combustion chamber <NUM>.

The other side surface of the engine body <NUM> is connected to an exhaust passage <NUM>. The exhaust passage <NUM> communicates with the exhaust ports <NUM> of each cylinder <NUM> and with the combustion chamber <NUM> via the exhaust ports <NUM>. Through the exhaust passage <NUM>, exhaust gas discharged from the combustion chamber <NUM> flows. Although not shown in detail, an upstream part of the exhaust passage <NUM> forms independent passages that branch off for the respective cylinders <NUM>. The upstream end of each independent passage is connected to the exhaust ports <NUM> of associated one of the cylinders <NUM>.

The exhaust passage <NUM> is provided with an exhaust gas purification systems having a plurality of (two in the example shown in <FIG>) catalyst converters. Although not shown, an upstream catalyst converter is located inside an engine compartment. This upstream catalyst converter includes a three-way catalyst <NUM> and a gasoline particulate filter (GPF) <NUM>. On the other hand, a downstream catalyst converter is located outside the engine compartment. This downstream catalyst converter includes a three-way catalyst <NUM>.

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 <NUM> may be omitted. The catalyst converters are not limited to the three-way catalysts <NUM> and <NUM>. The order of the three-way catalysts <NUM> and <NUM> and the GPF <NUM> may be changed as appropriate.

An EGR passage <NUM> constituting an external EGR system is interposed between the intake passage <NUM> and the exhaust passage <NUM>. The EGR passage <NUM> is for returning a part of the burned gas into the intake passage <NUM> and connects the intake passage <NUM> and the exhaust passage <NUM> together. The upstream end of the EGR passage <NUM> is connected between the upstream and downstream catalyst converters in the exhaust passage <NUM>. On the other hand, a downstream end of the EGR passage <NUM> is connected to the upstream side of the supercharger <NUM> in the intake passage <NUM>. The external EGR system is what is called a "low-pressure EGR system".

The EGR passage <NUM> includes a water-cooling EGR cooler <NUM>. The EGR cooler <NUM> cools the burned gas. The EGR passage <NUM> also includes an EGR valve <NUM>. The EGR valve <NUM> adjusts the flow rate of the burned gas flowing through the EGR passage <NUM>. 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 <NUM>.

In this exemplary configuration, an EGR system <NUM> includes the external EGR system including the EGR passage <NUM> and the EGR valve <NUM>, and the internal EGR system including the intake and exhaust electric S-VTs <NUM> and <NUM> described above.

The engine system includes the ECU <NUM> for operating the engine <NUM>. The ECU <NUM> is a controller including a known microcomputer as a base element. As shown in <FIG>, the ECU <NUM> includes a central processing unit (CPU) <NUM>, a memory <NUM> such as a random-access memory (RAM) and a read-only memory (ROM), and an input and output (I/O) bus <NUM>. The CPU <NUM> executes programs. The memory <NUM> stores the programs and data. The I/O bus <NUM> receives and outputs electrical signals.

This ECU <NUM> is connected to the injectors <NUM> described above, the spark plugs <NUM>, the intake electric S-VT <NUM>, the exhaust electric S-VT <NUM>, the fuel supply system <NUM>, the throttle valve <NUM>, the EGR valve <NUM>, the electromagnetic clutch <NUM> of the supercharger <NUM>, the air bypass valve <NUM>, and the swirl control valve <NUM>. As shown in <FIG> and <FIG>, the ECU <NUM> is also connected to various types of sensors SW1 to SW16. The sensors SW1 to SW16 output detection signals to the ECU <NUM>.

The sensors includes the following. An airflow sensor SW1 and a first intake air temperature sensor SW2 are arranged downstream of the air cleaner <NUM> in the intake passage <NUM>. A first pressure sensor SW3 is located downstream of the part of the intake passage <NUM> connected to the EGR passage <NUM> and upstream of the supercharger <NUM>. A second intake air temperature sensor SW4 is located downstream of the supercharger <NUM> in the intake passage <NUM> and upstream of the part of the intake passage <NUM> connected to the bypass passage <NUM>. A second pressure sensor SW5 is attached to the surge tank <NUM>. A pressure indicating sensor SW6 is attached to the cylinder head <NUM> in association with each cylinder <NUM>. An exhaust gas temperature sensor SW7 is located in the exhaust passage <NUM>.

The airflow sensor SW1 detects the flow rate of the fresh air flowing through the intake passage <NUM>. The first intake air temperature sensor SW2 detects the temperature of the fresh air flowing through the intake passage <NUM>. The first pressure sensor SW3 detects the pressure of the gas flowing into the supercharger <NUM>. The second intake air temperature sensor SW4 detects the temperature of the gas flowing out from the supercharger <NUM>. The second pressure sensor SW5 detects the pressure of the gas downstream of the supercharger <NUM>. The indicator sensor SW6 detects the pressures inside the combustion chamber <NUM>. The exhaust gas temperature sensor SW7 detects the temperature of the exhaust gas discharged from the combustion chamber <NUM>.

The sensors further include the following. A linear O<NUM> sensor SW8 is disposed upstream of the upstream catalyst converter in the exhaust passage <NUM>. A lambda O<NUM> sensor SW9 is disposed downstream of the three-way catalyst <NUM> in the upstream converter. A water temperature sensor SW10, a crank angle sensor SW11, an intake cam angle sensor SW12, and an exhaust cam angle sensor SW13 are attached to the engine body <NUM>. An accelerator position sensor SW14 is attached to an accelerator pedal mechanism. An EGR differential pressure sensor SW15 is disposed in the EGR passage <NUM>. A fuel pressure sensor SW16 is attached to the common rail <NUM> of the fuel supply system <NUM>.

The linear O<NUM> sensor SW8 and the lambda O<NUM> sensor SW9 each detect the oxygen concentration in the exhaust gas. The water temperature sensor SW10 detects the temperature of the coolant. The crank angle sensor SW11 detects the rotation angle of the crankshaft <NUM>. The intake cam angle sensor SW12 detects the rotation angle of the intake camshaft. The exhaust cam angle sensor SW13 detects the rotation angle of the exhaust camshaft. The accelerator position sensor SW14 detects the accelerator position. The EGR differential pressure sensor SW15 detects the differential pressure between the upstream and downstream sides of the EGR valve <NUM>. The fuel pressure sensor SW16 detects the pressure of the fuel to be supplied to the injectors <NUM>.

Based on the detection signals of these sensors, the ECU <NUM> determines the operating state of the engine <NUM> and calculates the control amounts of the devices. The ECU <NUM> outputs control signals related to the calculated control amounts to the injectors <NUM>, the spark plugs <NUM>, the intake electric S-VT <NUM>, the exhaust electric S-VT <NUM>, the fuel supply system <NUM>, the throttle valve <NUM>, the EGR valve <NUM>, the electromagnetic clutch <NUM> of the supercharger <NUM>, the air bypass valve <NUM>, and the swirl control valve <NUM>.

For example, the ECU <NUM> sets the target torque of the engine <NUM> and determines the target supercharging pressure based on the detection signal of the accelerator position sensor SW12 and a map set in advance. Then, the ECU <NUM> adjusts the opening degree of the air bypass valve <NUM> based on the target supercharging pressure and the differential pressure before and after the supercharger <NUM> obtained from the detection signals of the first pressure sensor SW3 and the second pressure sensor SW5. Accordingly, feedback control is performed so that the supercharging pressure reaches the target supercharging pressure.

The ECU <NUM> sets the target EGR rate, that is, the ratio of the EGR gas to the entire gas in the combustion chamber <NUM>, based on the operating state of the engine <NUM> and the map set in advance. Then, the ECU <NUM> 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 SW12. The ECU <NUM> adjusts the opening degree of the EGR valve <NUM> based on the differential pressure before and after the EGR valve <NUM> obtained from the detection signal of the EGR differential pressure sensor SW15. Through the determination and the adjustment, the ECU <NUM> performs feedback control so that the amount of the external EGR gas to be introduced into the combustion chamber <NUM> reaches the target amount of EGR gas.

The ECU <NUM> further executes feedback control of the air-fuel ratio upon satisfaction of predetermined control conditions. Specifically, the ECU <NUM> adjusts the amount of fuel injection by the injectors <NUM> based on the oxygen concentration in the exhaust gas detected by the linear O<NUM> sensor SW8 and the lambda O<NUM> sensor SW9 so that the air-fuel ratio of the air-fuel mixture reaches a desired value.

Details of the control of the engine <NUM> by the ECU <NUM> will be described later.

<FIG> illustrates operating range maps <NUM> and <NUM> of the warmed-up engine <NUM>. The operating range maps <NUM> and <NUM> of the engine <NUM> are defined by the load and speed of the engine <NUM>, and are divided into two ranges based on the magnitude of the speed of the engine <NUM>.

Specifically, the two ranges are: an SPCCI range (<NUM>) at a lower speed, specifically, at an engine speed lower than N1; and a CI range (<NUM>) at a higher speed, specifically, at an engine speed higher than or equal to N1. The SPCCI range (<NUM>) may here include low and medium speeds, if the entire operating range of the engine <NUM> is divided into three of low, medium, and high speed ranges in the direction of the speed. On the other hand, the CI range (<NUM>) may include the high speed range. The speed N1 may be about <NUM> rpm, for example.

In <FIG>, for easier understanding, each of the operating range maps <NUM> and <NUM> of the engine <NUM> is divided into two ranges. The map <NUM> shows the conditions of the air-fuel mixture and the combustion modes in operating states <NUM> to <NUM> of the engine <NUM> and driving and non-driving ranges of the supercharger <NUM>. The map <NUM> shows the opening degree of the swirl control valve <NUM> in each range. Note that the two-dot chain lines in <FIG> represent road-load lines of the engine <NUM>.

The engine <NUM> performs combustion by compressed autoignition for the main purpose of improving the fuel efficiency and the exhaust gas performance. More specifically, the engine <NUM> performs the SPCCI combustion described above in the SPCCI range (<NUM>). In the CI range (<NUM>), the engine <NUM> performs the CI combustion. Now, the operation of the engine <NUM> in the operating states <NUM> to <NUM> shown in <FIG> will be described in detail with reference to the fuel injection and ignition times shown in <FIG>. In <FIG>, the horizontal axis indicates the crank angle which advances from the left to the right in the drawing plane of <FIG>.

While operating in the SPCCI range (<NUM>), the engine <NUM> 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 <NUM> before the start of the compression. In the SPCCI combustion, the spark plug <NUM> forcibly ignites the air-fuel mixture in the combustion chamber <NUM> 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 <NUM>. The increase in the temperature in the combustion chamber <NUM> 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 <NUM> before the start of the compression. That is, even if the temperature in the combustion chamber <NUM> 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>, the reference character <NUM> denotes an example including fuel injection times (reference characters <NUM> and <NUM>), an ignition time (reference character <NUM>), and a combustion waveform (reference character <NUM>) in a low load operating mode <NUM> of the engine <NUM> in the SPCCI range (<NUM>). The combustion waveform represents a change in the heat generation rate with respect to the crank angle.

In the SPCCI combustion, the spark plug <NUM> ignites the air-fuel mixture at a predetermined timing near the compression top dead center (TDC on the right of <FIG>). 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 gentle slope at the rising. Although not shown, the pressure fluctuation (dp/dθ) in the combustion chamber <NUM> 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 <NUM>, the autoignition of the unburned air-fuel mixture occurs. In the example of <FIG>, at the autoignition timing, the waveform of the heat generation rate changes from the gentler slope to a steeper slope (see the reference character <NUM>). 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 <NUM> 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 the 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 the compression top dead center in the expansion stroke. Therefore, the SPCCI combustion is more advantageous in improving the fuel efficiency of the engine <NUM> than the SI combustion.

The EGR system <NUM> introduces the EGR gas into the combustion chamber <NUM> at a low load of the engine <NUM> in the SPCCI range (<NUM>) to improve the fuel efficiency of the engine <NUM>.

Specifically, a positive overlap period, in which both the intake and exhaust valves <NUM> and <NUM> 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 <NUM> to the intake and exhaust ports <NUM> and <NUM> returns and is reintroduced into the combustion chamber <NUM>. The internal EGR introduces hot burned gas (i.e., internal EGR gas) into the combustion chamber <NUM> and thus increases the temperature in the combustion chamber <NUM>, which is advantageous in stabilizing the SPCCI combustion.

At a low load of the engine <NUM>, the EGR valve <NUM> is fully closed. The external EGR gas is not introduced into the combustion chamber <NUM>.

The supercharger <NUM> is turned off at a low load of the engine <NUM> in the SPCCI range (<NUM>). Specifically, the supercharger <NUM> is turned off (see S/C OFF) at a lower speed and at low and medium loads in the SPCCI range (<NUM>). Even at the low and medium loads of the engine <NUM>, the supercharger <NUM> is turned on (see S/C ON) at a higher speed of the engine <NUM> and increases the supercharging pressure to ensure a required filling amount of the intake air.

When the supercharger <NUM> is turned off not to supercharge the gas in the intake passage <NUM>, the pressure inside the intake passage <NUM> is relatively low. The internal EGR gas is thus introduced into the combustion chamber <NUM> in the positive overlap period as described above.

When the supercharger <NUM> is turned on to supercharge the gas in the intake passage <NUM>, the pressure inside the intake passage <NUM> is relatively high. The gas in the intake passage <NUM> thus passes through the combustion chamber <NUM> of the engine body <NUM> and blows to the exhaust passage <NUM> in the positive overlap period. Accordingly, the burned gas remaining in the combustion chamber <NUM> is pushed out to the exhaust passage <NUM> and scavenged.

In an operation of the engine <NUM> in the SPCCI range (<NUM>), the swirl control valve <NUM> is fully closed or at a predetermined closing angle. Accordingly, a relatively strong swirl flow occurs in the combustion chamber <NUM>. The swirl flow is stronger on the periphery of the combustion chamber <NUM> and weaker at the center. As described above, each intake port <NUM> is the tumble port. Thus, an oblique swirl flow with a tumble component and a swirl component occurs in the combustion chamber <NUM>.

At a low load of the engine <NUM>, the swirl ratio is <NUM> 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>.

The tester shown in <FIG> is configured as follows. The cylinder head <NUM> is placed upside down on a base to connect the intake ports <NUM> to an intake air supplier (not shown). On the other hand, a cylinder <NUM> is placed on the cylinder head <NUM> to connect an impulse meter <NUM> including a honeycomb rotor <NUM> at the upper end of the cylinder <NUM>. The lower surface of the impulse meter <NUM> is located at a distance of <NUM>. 75D from the mating surface between the cylinder head <NUM> and the cylinder block. The term "D" means here the diameter of the cylinder bore. The tester measures, using the impulse meter <NUM>, the torque acting on the honeycomb rotor <NUM> due to the swirl flow (see the arrow in <FIG>) generated in the cylinder <NUM> 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> shows a relationship between the opening degree of the swirl control valve <NUM> and the swirl ratio of the engine <NUM>. In <FIG>, the opening degree of the swirl control valve <NUM> is expressed by the opening rate of the secondary passage <NUM> with respect to its fully opened cross-section. When the swirl control valve <NUM> is fully closed, the opening rate of the secondary passage <NUM> is <NUM>%. With an increase in the opening degree of the swirl control valve <NUM>, the opening rate of the secondary passage <NUM> becomes larger than <NUM>%. When the swirl control valve <NUM> is fully open, the opening rate of the secondary passage <NUM> is <NUM>%.

As illustrated in <FIG>, when the swirl control valve <NUM> is fully closed, the swirl ratio of the engine <NUM> is about <NUM>. At a low load of the engine <NUM> in the SPCCI range (<NUM>), the swirl ratio may range from <NUM> to <NUM>, both inclusive. The opening degree of the swirl control valve <NUM> may be adjusted to the opening rate ranging from <NUM>% to <NUM>%, 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 <NUM> at a low load of the engine <NUM> in the SPCCI range (<NUM>). That is, the excessive air ratio λ of the air-fuel mixture in the entire combustion chamber <NUM> is more than <NUM> (λ > <NUM>). More specifically, the A/F of the air-fuel mixture is <NUM> or more in the entire combustion chamber <NUM>. This ratio allows for reduction in RawNOX and improvement in the exhaust gas performance.

In the SPCCI range (<NUM>), in the low load operating mode <NUM> of the engine <NUM>, the air-fuel mixture is stratified between the center and periphery of the combustion chamber <NUM>. At the center of the combustion chamber <NUM>, the spark plug <NUM> is disposed. The periphery of the combustion chamber <NUM> is around the center and in contact with the liner of the cylinder <NUM>. The center of the combustion chamber <NUM> may be defined as a region with a weaker swirl flow, whereas the periphery of the combustion chamber <NUM> 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 <NUM> is higher than that on the periphery of the combustion chamber <NUM>. Specifically, the A/F of the air-fuel mixture at the center of the combustion chamber <NUM> ranges from <NUM> to <NUM>, whereas the A/F of the air-fuel mixture on the periphery of the combustion chamber <NUM> is <NUM> 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 <NUM> in the SPCCI range (<NUM>), the injector <NUM> injects the fuel into the combustion chamber <NUM> a plurality of times in the compression stroke (see the reference characters <NUM> and <NUM> in <FIG>). 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 here those 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 <NUM> by the ignition time to generate air-fuel mixture at the center and on the periphery of the combustion chamber <NUM>. 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 <NUM> at the center of the combustion chamber <NUM> without being diffused much. The fuel forms, together with a part of the fuel injected at the middle stage of the compression stroke, air-fuel mixture at the center of the combustion chamber <NUM>. As described above, the air-fuel mixture is stratified at the center and on the periphery of the combustion chamber <NUM>.

After the end of the fuel injection, the spark plug <NUM> ignites the air-fuel mixture at the center of the combustion chamber <NUM> at predetermined timing before the compression top dead center (see the reference character <NUM>). At this time, the air-fuel mixture contains the fuel at a relatively high concentration at the center of the combustion chamber <NUM>, 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 <NUM> in the SPCCI range (<NUM>), it is possible to achieve both the reduction in the combustion noise, and the improvement in the fuel efficiency due to a shorter combustion period.

In <FIG>, the reference character <NUM> denotes an example including fuel injection times (reference characters <NUM> and <NUM>), an ignition time (reference character <NUM>), and a combustion waveform (reference character <NUM>), in a medium load operation of the engine <NUM> in the SPCCI range (<NUM>).

The EGR system <NUM> introduces the EGR gas into each combustion chamber <NUM> in a medium load operation of the engine <NUM> as in a low load operation. Specifically, in a low load and a lower speed operation of the engine <NUM> within the medium load range, the positive overlap period, in which both the intake and exhaust valves <NUM> and <NUM> 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 <NUM> to the intake and exhaust ports <NUM> and <NUM> returns and is reintroduced into the combustion chamber <NUM>. That is, the internal EGR gas is introduced into the combustion chamber <NUM>.

In a high load or a high speed operation of the engine <NUM> within the medium load range, the supercharger <NUM> 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 <NUM> is turned on to supercharge the gas in the intake passage <NUM>, the pressure inside the intake passage <NUM> is relatively high. The residual gas (i.e., hot burned gas) in the combustion chamber <NUM> is thus scavenged in the positive overlap period as described above.

In a medium load operation of the engine <NUM>, external EGR is performed in which the exhaust gas cooled by the EGR cooler <NUM> is introduced through the EGR passage <NUM> into the combustion chamber <NUM>. That is, the external EGR gas with a lower temperature than that of the internal EGR gas is introduced into the combustion chamber <NUM>. The introduction of at least one of the internal EGR gas and the external EGR gas into the combustion chamber <NUM> leads to adjustment of the temperature in the combustion chamber <NUM> to an appropriate temperature. Note that the EGR rate increases with an increase in the load on the engine <NUM>.

In a medium load operation of the engine <NUM>, the swirl control valve <NUM> 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 <NUM> or more occurs in the combustion chamber <NUM>. With an increase in the strength of the swirl flow, the turbulent energy inside the combustion chamber <NUM> 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, the combustion noise decreases and the 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 = <NUM>) in the entire combustion chamber <NUM> in a medium load operation of the engine <NUM>. At the stoichiometric air-fuel ratio, a three-way catalyst purifies the exhaust gas discharged from the combustion chamber <NUM> so that the engine <NUM> 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 <NUM> ± <NUM>.

In a medium load operation of the engine <NUM>, the injector <NUM> injects the fuel into the combustion chamber <NUM> separately in the intake and compression strokes (see the reference characters <NUM> and <NUM> in <FIG>). Specifically, the injector performs first injection <NUM> to inject the fuel in the period from the middle stage to final stage of the intake stroke, and second injection <NUM> to inject the fuel at a second half of the compression stroke. The middle stage and final stage of the intake stroke may be here those 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, the first and second halves of the compression stroke may be those 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 <NUM>, the fuel is injected at a time away from the ignition time. At the time of the injection, the piston <NUM> is away from the top dead center, the fuel reaches a squish area <NUM> outside the cavity <NUM> and is distributed substantially evenly in the combustion chamber <NUM> to generate air-fuel mixture. In the second injection <NUM>, the fuel is injected when the piston <NUM> is closer to the compression top dead center and thus enters the cavity <NUM> to generate air-fuel mixture in the area inside of the cavity <NUM>.

In accordance with the second injection <NUM> for injecting the fuel into the cavity <NUM>, the gas flow occurs in the area inside the cavity <NUM>. If a longer time is required to the ignition, the turbulent energy in the combustion chamber <NUM> weakens with the progress of the compression stroke. However, the second injection <NUM> is closer to the ignition time than the first injection <NUM>, which allows the spark plug <NUM> to ignite the air-fuel mixture in the area inside the cavity <NUM> while maintaining the large turbulent energy in the cavity <NUM>. 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 <NUM> at the second half of the compression stroke, latent heat of fuel vaporization reduces the temperature in the combustion chamber <NUM> to reduce induction of abnormal combustion such as preignition or knocking. The fuel injected in the second injection <NUM> can be stably burned by the flame propagation. The ratio between the amount of the first injection <NUM> and the amount of the second injection <NUM> may be <NUM>:<NUM>, for example.

In the combustion chamber <NUM>, a substantially homogeneous air-fuel mixture with an excessive air ratio λ of <NUM> ± <NUM> as a whole is generated by the injector <NUM> performing the first and second injections <NUM> and <NUM>. Since the air-fuel mixture is substantially homogeneous, the fuel efficiency improves with a decrease in unburned loss and the exhaust gas performance improves with a decrease in smoke (soot).

The spark plug <NUM> ignites the air-fuel mixture at a predetermined timing before the compression top dead center (see the reference character <NUM>), 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 <NUM> is mainly subjected to the SI combustion. The fuel injected in the first injection <NUM> is mainly subjected to the CI combustion.

In <FIG>, reference character <NUM> denotes an example including a fuel injection time (reference character <NUM>), an ignition time (reference character <NUM>), and a combustion waveform (reference character <NUM>) in a high load operation of the engine <NUM> within the SPCCI range (<NUM>).

The EGR system <NUM> also introduces the EGR gas into each combustion chamber <NUM> in a high load operation of the engine <NUM>.

Specifically, in a high load operation of the engine <NUM>, the external EGR is performed in which the exhaust gas cooled by the EGR cooler <NUM> is introduced through the EGR passage <NUM> into the combustion chamber <NUM>. The EGR rate continuously increases with an increase in the load on the engine <NUM> within the medium and high load ranges of in the SPCCI range (<NUM>). The introduction of the external EGR gas cooled by the EGR cooler <NUM> into the combustion chamber <NUM> leads to adjustment of the temperature in the combustion chamber <NUM> to an appropriate temperature and reduction in induction of abnormal combustion, such as preignition or knocking, of the air-fuel mixture.

At a load on the engine <NUM> closer to the full load, there is a need to increase the amount of fresh air to be introduced into the combustion chamber <NUM> to cope with an increase in the amount of fuel. For this purpose, at a load of the engine <NUM> closer to the full load in the SPCCI range (<NUM>), the EGR rate of the external EGR may decrease.

In a high load operation of the engine <NUM> as well, the positive overlap period, in which both the intake and exhaust valves <NUM> and <NUM> are open, is provided near the exhaust top dead center.

In a high load operation of the engine <NUM> as well, the supercharger <NUM> 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 <NUM> in the positive overlap period.

In a high load operation of the engine <NUM> as well, the swirl control valve <NUM> is fully closed or at a predetermined closing angle. Accordingly, a strong swirl flow with a swirl ratio of <NUM> or more occurs in the combustion chamber <NUM>.

In a high load operation of the engine <NUM>, 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 ≤ <NUM>) in the entire combustion chamber <NUM>.

In a high load operation mode <NUM> of the engine <NUM>, the injector <NUM> starts injecting the fuel in the intake stroke (see the reference character <NUM> in <FIG>). Specifically, the fuel injection <NUM> may start injecting the fuel at <NUM>°CA before the compression top dead center. The fuel injection <NUM> may continue over the intake stroke and end in the compression stroke. The fuel injection <NUM> may start at the first half of the intake stroke, which allows the fuel spray to hit the opening edge of the cavity <NUM>. A part of the fuel enters the squish area <NUM> of the combustion chamber <NUM>, that is, the area outside the cavity <NUM> (see <FIG>), whereas the rest enters the area inside the cavity <NUM>. At this time, the swirl flow is stronger on the periphery of the combustion chamber <NUM>, and weaker at the center of the combustion chamber <NUM>. Accordingly, the fuel that has entered the area inside the cavity <NUM> 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 air-fuel mixture for the CI combustion on the periphery of the combustion chamber <NUM>. 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 air-fuel mixture for the SI combustion at the center of the combustion chamber <NUM>.

In a high load operation of the engine <NUM>, the fuel concentration of the air-fuel mixture is set to be higher on the periphery of the combustion chamber <NUM> 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 <NUM> than at the center.

Specifically, the excessive air ratio λ of the air-fuel mixture is <NUM> or lower in one preferred embodiment at the center of the combustion chamber <NUM>, and <NUM> or lower and, in one preferred embodiment, lower than <NUM> on the periphery of the combustion chamber <NUM>. At the center of the combustion chamber <NUM>, the air-fuel ratio (A/F) of the air-fuel mixture may range from <NUM> to the stoichiometric air-fuel ratio (i.e., <NUM>), for example. Alternatively, at the center of the combustion chamber <NUM>, 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 <NUM>, the air-fuel ratio of the air-fuel mixture may range from <NUM> to the stoichiometric air-fuel ratio, for example, and may range from <NUM> to <NUM> in one preferred embodiment. The excessive air ratio λ less than <NUM> on the periphery of the combustion chamber <NUM> 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 <NUM>, the air-fuel ratio of the air-fuel mixture may range from <NUM> to the stoichiometric air-fuel ratio, for example, and may range from <NUM> to <NUM> in one preferred embodiment.

Near the compression top dead center, the spark plug <NUM> ignites the air-fuel mixture in the combustion chamber <NUM> (see the reference character <NUM>). The spark plug <NUM> may perform the ignition, for example, after the compression top dead center. Since the spark plug <NUM> is located at the center of the combustion chamber <NUM>, the ignition of the spark plug <NUM> 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 <NUM>, the flame stably propagates after the ignition of the spark plug <NUM> in the SPCCI combustion.

With an increase in the load of the engine <NUM>, the amount of the fuel injection and the temperature in the combustion chamber <NUM> increase, which creates a condition easily starting the CI combustion earlier. That is, a higher load of the engine <NUM> tends to cause abnormal combustion, such as preignition and knocking, of the air-fuel mixture. However, as described above, the temperature on the periphery of the combustion chamber <NUM> 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 <NUM> in the SPCCI combustion, the stratification of the air-fuel mixture inside the combustion chamber <NUM> and the generation of a strong swirl flow inside the combustion chamber <NUM> allow sufficient SI combustion before the start of the CI combustion. This results in reduction in the combustion noise and an excessive increase in the 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 <NUM> makes the CI combustion more moderate, which is advantageous in reducing the combustion noise. In addition, since the CI combustion requires a shorter combustion period, the torque and the thermal efficiency improve at a high load of the engine <NUM>. Therefore, in the engine <NUM>, the SPCCI combustion is performed within a high load range, thereby improving the fuel efficiency while reducing the combustion noise.

At a high speed of the engine <NUM>, a shorter time is required for the crank angle to change by <NUM>°. For example, in the high speed and high load range, it is thus difficult to stratify the air-fuel mixture in the combustion chamber <NUM> to perform the SPCCI combustion. On the other hand, since the engine <NUM> 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 <NUM> performs a new CI combustion mode in a higher speed operation in the CI range (<NUM>). The CI range (<NUM>) 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 <NUM> are as follows. A lean air-fuel mixture and/or a strong flow in the combustion chamber <NUM> do(es) not allow the progress of the combustion by the flame propagation, even after the spark plug <NUM> has ignited the air-fuel mixture. The combustion mode in the CI range (<NUM>) is based on the following new finding obtained when the air-fuel mixture was viewed microscopically. If the spark plug <NUM> 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>, reference character <NUM> denotes an example including fuel injection times (reference characters <NUM> and <NUM>), an ignition time (reference character <NUM>), and a combustion waveform (reference character <NUM>), in a high load operating state <NUM> of the engine <NUM> in the SPCCI range (<NUM>).

The air-fuel ratio (A/F) of the air-fuel mixture is basically equal to the stoichiometric air-fuel ratio (i.e., A/F = <NUM>) in the entire combustion chamber <NUM> in an operation of the engine <NUM> in the CI range (<NUM>). The excessive air ratio λ of the air-fuel mixture may be <NUM> ± <NUM>. In a high load range including the all loads within the CI range (<NUM>), the excessive air ratio λ of the air-fuel mixture may be lower than <NUM>.

As shown in the map <NUM> in <FIG>, in an operation of the engine <NUM> in the CI range (<NUM>), the supercharger <NUM> is turned on (see S/C ON) throughout all the ranges of the engine <NUM> to increase the supercharging pressure.

<FIG> illustrates a change in the flow strength in the combustion chamber <NUM> from the intake stroke to the compression stroke. In an operation of the engine <NUM> in the CI range (<NUM>), the swirl control valve <NUM> is fully open as shown in the map <NUM> in <FIG>. Accordingly, no swirl flow but only a tumble flow occurs in the combustion chamber <NUM>. Such full opening of the swirl control valve <NUM> improves the filling efficiency, while reducing pump losses at a high speed of the engine <NUM>.

As the intake air flows into the combustion chamber <NUM> in the intake stroke, a tumble flow occurs and the flow in the combustion chamber <NUM> gradually strengthens. The flow in the combustion chamber <NUM> that has strengthened in the intake stroke weakens once at the post-stage of the intake stroke. However, with a lift of the piston <NUM> toward the top dead center in the compression stroke, the flow in the combustion chamber <NUM> strengthens again due to what is called a "spin-up phenomenon". As indicated by the white arrow in <FIG>, the flow in the combustion chamber <NUM> 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 <NUM> 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 <NUM>. The axis is parallel to the crankshaft <NUM> passing through the center of gravity of the combustion chamber <NUM>, whose position varies depending on a change in the volume of the combustion chamber <NUM>. The angular velocity ω of the intake air can be obtained as follows. Specifically, the inside of the combustion chamber <NUM> 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 air-fuel mixture with a low fuel concentration with respect to the stoichiometric air-fuel ratio in the combustion chamber <NUM> in or before the specific period does not allow combustion by the flame propagation, even after the spark plug <NUM> ignites the air-fuel mixture in the specific period. That is, in the CI range (<NUM>), the spark plug <NUM> ignites the air-fuel mixture when the conditions inside the combustion chamber <NUM> fall within the broken reaction zone. Accordingly, the flame is stored in the combustion chamber <NUM> without starting the combustion by the flame propagation.

After that, as the crank angle advances, the flow in the combustion chamber <NUM> weakens. In addition, the fuel is additionally supplied into the combustion chamber <NUM> to increase the fuel concentration of the air-fuel mixture. Then, the conditions inside the combustion chamber <NUM> 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 <NUM>. 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>. In an operation of the engine <NUM> in the CI range (<NUM>), the injector <NUM> performs the fuel injection (i.e., the first fuel injection <NUM>) 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 <NUM>. 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 <NUM> is determined by the load of the engine <NUM> and the division ratio between the first fuel injection <NUM> and the second fuel injection <NUM> described later. Note that the fuel injection period varies depending on the amount of the first fuel injection <NUM>. The start of the first fuel injection <NUM> may be set as appropriate in accordance with the amount of the first fuel injection <NUM> so as to form air-fuel mixture with a low fuel concentration with respect to the stoichiometric air-fuel ratio in the combustion chamber <NUM> at least in or before the specific period described above.

After the end of the first fuel injection, the spark plug <NUM> ignites the air-fuel mixture (see the reference character <NUM>). An upper illustration <NUM> of <FIG> illustrates an ignition time in a high load operation of the engine <NUM> in the CI range (<NUM>). The speed N1 in <FIG> corresponds to the speed N1 of the map shown in <FIG>. In the upper illustration <NUM> of <FIG>, the vertical axis represents the crank angle, which advances toward the top of the vertical axis.

In the upper illustration <NUM> of <FIG>, the hatched range indicates the ignition time of the spark plug <NUM>. The spark plug <NUM> ignites the air-fuel mixture at an appropriate timing within the hatched range.

The spark plug <NUM> 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 <NUM> without allowing the progress of the combustion by the flame propagation. The flame is dispersed or diffused by the flow in the combustion chamber <NUM>. As is apparent from <FIG>, 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 <NUM> is weak. Thus, the ignition time has an advance limit (i.e., the upper line in the upper illustration <NUM>) and a retard limit (i.e., the lower line in the upper illustration <NUM>) so that the spark plug <NUM> ignites the air-fuel mixture, when the conditions inside the combustion chamber <NUM> fall within the broken reaction zone, as shown in the upper illustration <NUM> of <FIG>. The spark plug <NUM> 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 <NUM> in the CI range (<NUM>) is much more advanced than the 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 <NUM> of <FIG>. 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 <NUM>. In the example of the upper illustration <NUM>, the ignition time advances with an increase in the speed of the engine <NUM>. The variation in the advance limit with an increase in the speed of the engine <NUM> 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 <NUM>.

The spark plug <NUM> may perform a plurality of ignitions within the specific period. This increases the number of flames generated in the combustion chamber <NUM> and allows a strong flow in the combustion chamber <NUM> to diffuse the large number of flames into the combustion chamber <NUM>. 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 <NUM> has ignited the air-fuel mixture, the injector <NUM> injects the fuel into the combustion chamber <NUM> (i.e., the second fuel injection <NUM>). The fuel concentration of the air-fuel mixture in the combustion chamber <NUM> increases. The second fuel injection <NUM> 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 <NUM> weakens, whereby the conditions inside the combustion chamber <NUM> fall out of the broken reaction zone. In addition, the temperature and the pressure inside the combustion chamber <NUM> are increased by motoring as the piston comes closer to the compression top dead center. Accordingly, near the compression top dead center, the flame stored starts the combustion of the air-fuel mixture by the autoignition at once (see the reference character <NUM>). The center of gravity of this combustion is closer to the compression top dead center, which improves the thermal efficiency of the engine <NUM>. 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 <NUM>, whether or not the conditions inside the combustion chamber <NUM> fall out of the broken reaction zone changes. A change in the time when the conditions of the combustion chamber <NUM> 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 <NUM> may thus be adjusted as appropriate to start the combustion of the air-fuel mixture at an appropriate timing.

As illustrated in <FIG>, the second fuel injection <NUM> may be performed after the spark plug <NUM> has ignited the air-fuel mixture. A delay in the end of the fuel injection <NUM> shortens the vaporization time of the fuel injected in the second fuel injection <NUM>, which may be disadvantageous in the exhaust gas emission performance or the fuel efficiency. The start of the second fuel injection <NUM> may be set as appropriate based on the amount of the second fuel injection <NUM> not to cause a delay of the injection end of the second fuel injection <NUM>. For example, the lower illustration <NUM> of <FIG> illustrates the start of the second fuel injection <NUM>. In the lower illustration <NUM> of <FIG>, the vertical axis of represents the crank angle, which advances toward the top of the vertical axis. In the lower illustration <NUM> of the figure, the hatched area represents the start of the second fuel injection <NUM>. The injector <NUM> performs the second fuel injection <NUM> at an appropriate timing within the hatched range.

With a change in the start of the second fuel injection <NUM>, the time when the conditions inside the combustion chamber <NUM> fall out of the broken reaction zone changes. Thus, with a change in the start of the second fuel injection <NUM> by the injector <NUM>, the start of the combustion of the air-fuel mixture changes. The start of the second fuel injection <NUM> may be adjusted to start the combustion of the air-fuel mixture at an appropriate timing.

The start of the second fuel injection <NUM> may change in accordance with the magnitude of the speed of the engine <NUM>. Specifically, with an increase in the speed of the engine <NUM>, the ignition timing advances. In accordance with the advance, the start of the second fuel injection <NUM> may advance. In the example of the lower illustration <NUM>, the start of the second injection advances with an increase in the speed of the engine <NUM>. The variation in the advance limit (i.e., the slope of the upper line in the lower illustration <NUM>) with an increase in the speed of the engine <NUM> is larger than the variation in the retard limit (i.e., the slope of the lower line in the lower illustration <NUM>).

The ratio between the amounts of the first and second injection <NUM> and <NUM> may be set as appropriate.

The EGR system <NUM> introduces the external and/or internal EGR gas into the combustion chamber <NUM> in an operation of the engine <NUM> in the CI range (<NUM>). The EGR rate in the combustion chamber <NUM> may be adjusted to start the combustion of the air-fuel mixture at an appropriate timing.

In the CI range (<NUM>), water may be injected into the combustion chamber <NUM> at an appropriate timing in the compression stroke to start the combustion of the air-fuel mixture at an appropriate timing.

In the maps <NUM> and <NUM> shown in <FIG>, the combustion mode described above may be employed throughout the entire CI range (<NUM>) or the load range on the R-L line in the CI area (<NUM>).

<FIG> illustrates a flowchart of the control of the engine <NUM> described above and executed by the ECU <NUM>.

First, in a step S1 after the start of the flow, the ECU <NUM> reads signals of the various sensors SW1 to SW16. In the subsequent step S2, the ECU <NUM> determines whether or not the operating state of the engine <NUM> is within the high speed range. The high speed range corresponds to the CI range (<NUM>) described above. In the step S2, the ECU <NUM> may determine whether or not the engine speed is equal to or higher than N1. If the determination in the step S2 is YES, the control process proceeds to a step S3. If the determination in the step S2 is NO, the process proceeds to a step S6.

Steps S3 to S5 are the control steps within the CI range (<NUM>). The steps S3 to S5 proceed in this order.

First, in the step S3, the ECU <NUM> executes the first fuel injection by the injector <NUM>. Accordingly, leaner air-fuel mixture with respect to the stoichiometric air-fuel ratio is generated in the combustion chamber <NUM>. In the subsequent step S4, the ECU <NUM> executes ignition by the spark plug <NUM> at a predetermined time. As described above, the spark plug <NUM> performs the ignition before the post-mid stage of the compression stroke. In the step S5, the ECU <NUM> executes then the second fuel injection by the injector <NUM>. As a result, near the compression top dead center, the flame made in advance in the combustion chamber <NUM> starts the combustion of the air-fuel mixture at once by autoignition.

On the other hand, in the step S6, the control in the SPCCI range (<NUM>) is performed in accordance with the load of the engine <NUM>.

Note that application of the disclosed technique is not limited to the engine <NUM> with the configuration described above. The engine may employ various configurations.

<FIG> shows a configuration of an engine <NUM> according to a variation. The engine <NUM> includes a turbocharger <NUM> in place of the mechanical supercharger <NUM>.

The turbocharger <NUM> includes a compressor <NUM> in the intake passage <NUM> and a turbine <NUM> in the exhaust passage <NUM>. The turbine <NUM> is rotated by the exhaust gas flowing through the exhaust passage <NUM>. The compressor <NUM> is rotated by the rotational drive of the turbine <NUM> and supercharges the gas in the intake passage <NUM> to be introduced into the combustion chamber <NUM>.

The exhaust passage <NUM> includes an exhaust bypass passage <NUM>. The exhaust bypass passage <NUM> connects parts of the exhaust passage <NUM> upstream and downstream of the turbine <NUM> as to bypass the turbine <NUM>. The exhaust bypass passage <NUM> includes a waste gate valve <NUM>. The waste gate valve <NUM> adjusts the flow rate of the exhaust gas flowing through the exhaust bypass passage <NUM>.

In this exemplary configuration, the turbocharger <NUM>, the bypass passage <NUM>, the air bypass valve <NUM>, the exhaust bypass passage <NUM>, and the waste gate valve <NUM> constitute a supercharging system <NUM> in the intake and exhaust passages <NUM> and <NUM>.

The engine <NUM> switches opening/closing of the air bypass valve <NUM> and the waste gate valve <NUM> to or not to cause the turbocharger <NUM> to supercharge the gas to be introduced into the combustion chamber <NUM>.

When the gas introduced into the combustion chamber <NUM> is not supercharged, the waste gate valve <NUM> opens. Accordingly, the gas flowing through the exhaust passage <NUM> bypasses the turbine <NUM>, that is, passes not through the turbine <NUM> but through the bypass passage <NUM> into the catalyst converters. Then, the turbine <NUM> does not receive the flow of the exhaust gas and thus does not drive the turbocharger <NUM>. At this time, the air bypass valve <NUM> is fully open. As a result, the gas flowing through the intake passage <NUM> flows through none of the compressor <NUM> or the intercooler <NUM> but through the bypass passage <NUM> into the surge tank <NUM>.

When the gas to be introduced into the combustion chamber <NUM> is supercharged, the waste gate valve <NUM> is not fully open but closed a little. Accordingly, at least a part of the exhaust gas flowing through the exhaust passage <NUM> passes through the turbine <NUM> and flows to the catalyst converters. Then, the turbine <NUM> rotates upon receipt of the exhaust gas and drives the turbo turbocharger <NUM>. When the turbocharger <NUM> is driven, the gas in the intake passage <NUM> is supercharged by the rotation of the compressor <NUM>. At this time, if the air bypass valve <NUM> is open, a part of the gas that has passed through the compressor <NUM> flows back from the surge tank <NUM> through the bypass passage <NUM> to the upstream side of the compressor <NUM>. The supercharging pressure of the gas inside the intake passage <NUM> can be controlled by adjusting the opening degree of the air bypass valve <NUM> as in the case using the mechanical supercharger described above.

Whether or not the turbocharger <NUM> supercharges the gas in the intake passage <NUM> may be determined in accordance with a map <NUM> shown in <FIG>, for example. That is, the turbocharger <NUM> may not perform the supercharging in the low load range in the SPCCI range (<NUM>)(see T/C OFF), whereas the turbocharger <NUM> may perform the supercharging in the medium and high load ranges in the SPCCI range (<NUM>) as well as in the CI range (<NUM>) (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 <NUM> and <NUM> at the activation temperatures, the waste gate valve <NUM> may open to bypass the turbine <NUM>, thereby reducing heat dissipation in the turbine <NUM> and supplying hot exhaust gas to the three-way catalysts <NUM> and <NUM>.

The operation of the engine <NUM> including the turbocharger <NUM> may be controlled in accordance with the flowchart shown in <FIG>. The engine <NUM> also can improve the thermal efficiency, while reducing abnormal combustion at a high speed of the engine <NUM>.

Although not shown, the disclosed technology is applicable to a natural intake engine without any supercharger.

In the configuration described above, the first fuel injection <NUM> generates homogeneous or substantially homogeneous air-fuel mixture with the A/F higher than the stoichiometric air-fuel ratio in the combustion chamber <NUM>. The time of the first fuel injection <NUM> may be adjusted to locally generate air-fuel mixture with an A/F higher than the stoichiometric air-fuel ratio near the spark plug <NUM> at the ignition timing.

In addition to the injector for directly injecting the fuel into the combustion chamber <NUM>, 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.

Claim 1:
An engine (<NUM>,<NUM>) control method of executing a cycle including an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke in a combustion chamber (<NUM>), the method comprising:
a step of creating, using a tumble port, a strong tumble flow in the combustion chamber (<NUM>),
a first fuel supply step in which a fuel supply unit (<NUM>) supplies fuel into the combustion chamber (<NUM>) so as to generate an air-fuel mixture having an air-fuel mass ratio (A/F) higher than a stoichiometric air-fuel ratio;
an ignition step in which an ignition unit (<NUM>) arranged in the combustion chamber (<NUM>) makes flame after the supply of the fuel to the combustion chamber (<NUM>) in the first fuel supply step and at a timing when the tumble ratio in the combustion chamber (<NUM>) is equal to or higher than a predetermined value in a compression stroke during or before a post-mid stage such that the lean air-fuel mixture and the strong flow in the combustion chamber do not allow the progress of the combustion by flame propagation and the flame does not go out but is stored 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 a sequential manner,
after the ignition step, a second fuel supply step in which the fuel supply unit (<NUM>) supplies fuel into the combustion chamber (<NUM>) to increase a fuel concentration of the air-fuel mixture in the combustion chamber (<NUM>), and
a step in which combustion of the air-fuel mixture is started by the stored flame by auto-ignition after the second fuel supply step.