Compression ignition engine, control system therefor, method for controlling an engine and computer program product

Abstract

Compression ignition engine (1), comprising: a plurality of cylinders (11), each of which is equipped with a combustion chamber (17); a spark plug (25) which is arranged in each of the combustion chambers (17); a fuel injection device (6) arranged to be directed into each combustion chamber (17); and a control unit (10) which includes a processor (101) which is connected to the spark plug (25) and the fuel injection device (6) and is configured to output a control signal to the spark plug (25) and the fuel injection device (6), respectively, wherein When the engine (1) operates at a load above a given load, the control unit (10) causes the engine (1) to operate with all cylinders by supplying fuel to all or a plurality of cylinders (11), and when the engine (1) operates at a load below the given load, the control unit (10) causes the engine (1) to operate with a reduced number of cylinders by supplying fuel to one or some of the plurality of cylinders (11), and In operation with reduced cylinders, the control unit (10) regulates or controls the fuel injection device (6) to inject the fuel to one or some of the majority of cylinders (11) to produce a mixed gas, regulates or controls the spark plug (25) to ignite the mixed gas, and causes the engine (1) to initiate combustion by spark ignition (SI) at an air-fuel ratio greater than a stoichiometric air-fuel ratio and at a high compression ratio above a given value, in which the mixed gas is ignited by the spark plug (25) to burn by flame propagation, and then to carry out combustion by compression ignition (CI) in which unburned mixed gas ignites by auto-ignition. When the engine (1) operates under a load below the given load, the control unit (10) switches an operating mode of the engine (1) between a first combustion mode, in which SI combustion and CI combustion are carried out in that order at an air-fuel ratio substantially equal to a stoichiometric air-fuel ratio, and a second combustion mode, in which SI combustion and CI combustion are carried out in that order at an air-fuel ratio greater than the stoichiometric air-fuel ratio, and when the engine (1) is operating in the first combustion mode, operation with reduced cylinders is not carried out.

Description

TECHNICAL FIELD

[0001] The present disclosure relates to a compression ignition engine, a control system therefor, a method for controlling an engine, and a computer program product. BACKGROUND OF REVELATION

[0002] JP 2007-154859 A discloses a control device for a spark-ignition gasoline engine, including an HCCI (Homogeneous Charge Compression Ignition) executing part that performs compression self-ignition operation at least within a partial load operating range of the engine, and an ignition assisting part that assists ignition in compression self-ignition operation within the partial load operating range. On a low engine load side of the partial load operating range, the control device performs a reduced-cylinder operation. JP 2007-154859 A also discloses performing combustion at an excess air ratio λ=1 during the reduced-cylinder operation.

[0003] However, due to performing combustion at the excess air ratio λ=1 during the reduced-cylinder operation, the conventional control device for the spark-ignition gasoline engine provided with the ignition assisting part and the HCCI executing part causes a problem of poor fuel efficiency.

[0004] Although it can be thought of making an air-fuel ratio A / F lean within a low engine load operating range, this makes flame propagation difficult in SI (spark ignition) combustion and also fluctuates combustion timing in CI (compression ignition) combustion.

[0005] JP 2006-283 667 A discloses a control or regulating device for a spark-ignition multi-cylinder engine in which, in the partial load range, combusted gas is introduced from one cylinder in the exhaust stroke into a subsequent cylinder in the intake stroke. SUMMARY OF REVELATION

[0006] The present disclosure has been made with a view to solving the objects or problems of the above-described prior art and aims to improve fuel efficiency, perform stable SI combustion operation, and also perform subsequent CI combustion operation at a desired timing even during reduced-cylinder operation within a low engine load operation range.

[0007] SI combustion is combustion accompanied by flame propagation initiated by forced ignition of a mixed gas within a combustion chamber, and CI combustion is combustion initiated by the mixed gas inside the combustion chamber self-igniting through compression. In a combustion mode that combines SI combustion and CI combustion, the mixed gas inside the combustion chamber is forcedly ignited to initiate combustion through flame propagation. Then, the heat generated by SI combustion and the pressure raised by flame propagation cause the unburned mixed gas inside the combustion chamber to burn through compression ignition. Hereinafter, this combustion mode is referred to as SPCCI (Spark Controlled Combustion).controlled compression ignition) combustion.

[0008] In compression-ignition combustion, the compression ignition timing changes significantly as the temperature inside the combustion chamber varies before compression starts. In this regard, the change in the temperature inside the combustion chamber before compression starts can be reduced by adjusting the heat generation amount during SI combustion. For example, the compression ignition timing can be controlled by controlling the ignition timing to adjust the start timing of SI combustion according to the temperature inside the combustion chamber before compression starts.

[0009] SI combustion through flame propagation results in a relatively gentle pressure rise compared to CI combustion, thus reducing combustion noise. Furthermore, CI combustion shortens the combustion period compared to SI combustion, making SPCCI combustion advantageous for improving fuel efficiency.

[0010] When the engine operates within a low-load range, the fuel injection amount decreases, and the temperature inside the combustion chamber also decreases. Therefore, even if SPCCI combustion is to be performed, spark ignition is difficult to perform, and SI combustion becomes unstable. It may also be possible that CI combustion does not occur due to insufficient temperature.

[0011] Thus, the present inventors aim to achieve the stable SPCCI combustion while making an air-fuel ratio (A / F) lean even within such a low load range which is detrimental to the SPCCI combustion.

[0012] The teaching disclosed here is a configuration in which, within the low load range, a so-called reduced cylinder operation in which fuel is supplied only to one or some of a plurality of cylinders while not being supplied to the other cylinders under a certain condition is performed at a large compression ratio above a given value.

[0013] Specifically, according to one aspect of the present disclosure, a compression ignition engine includes a plurality of cylinders each formed with a combustion chamber, a spark plug disposed in each of the combustion chambers, a fuel injector disposed to face each combustion chamber, and a control unit connected to the spark plug and the fuel injector and configured to output a control signal to each of the spark plug and the fuel injector. When the engine is operating at a load above a given load, the control unit causes the engine to perform all-cylinder operation by supplying fuel to all of the plurality of cylinders, and when the engine is operating at a load below the given load, the control unit causes the engine to perform all-cylinder operation by supplying fuel to all of the plurality of cylinders.Control unit that the engine performs a reduced-cylinder operation by supplying the fuel to one or some of the plurality of cylinders. In the reduced-cylinder operation, the control unit controls the fuel injector to inject the fuel to one or some of the plurality of cylinders to generate a mixture gas, controls the spark plug to ignite the mixture gas, and causes the engine, at an air-fuel ratio greater than a stoichiometric air-fuel ratio and at a large compression ratio above a given value, to start combustion by spark ignition (SI) in which the mixture gas is ignited by the spark plug to generate a mixture by flame propagation.-propagation, and then perform compression ignition (CI) combustion in which unburned mixture gas is ignited by compression ignition. When the engine operates at a load below the given load, the control unit switches an operation mode of the engine between a first combustion mode in which SI combustion and CI combustion are performed in this order at an air-fuel ratio substantially equal to or equivalent to a stoichiometric air-fuel ratio, and a second combustion mode in which SI combustion and CI combustion are performed in this order at an air-fuel ratio greater than the stoichiometric air-fuel ratio. When the engine operates in the first combustion mode, reduced-cylinder operation is not performed.

[0014] Accordingly, in the first combustion mode in which the SI combustion and the subsequent CI combustion are performed at the air-fuel ratio equivalent to the stoichiometric air-fuel ratio, since the all-cylinder operation is performed instead of the reduced-cylinder operation, generation of NO x when switching the control system to operation with reduced cylinders.

[0015] Here, the “engine” can be a four-stroke engine, which operates by the combustion chamber repeating an intake stroke, compression stroke, expansion stroke and exhaust stroke.

[0016] According to this configuration, within a low load range (under the given load), reduced-cylinder operation is performed in which fuel is supplied to one or some of the plurality of cylinders, and SI combustion and subsequent CI combustion are performed at the air-fuel ratio greater than the stoichiometric air-fuel ratio (a lean A / F) and the large compression ratio above the given value (e.g., a geometric compression ratio is approximately 13:1 or higher). By performing reduced-cylinder operation while the A / F is lean, discharge of unprocessed NO x (raw NO x) and improves fuel efficiency. Furthermore, during reduced-cylinder operation, since CI combustion is performed at a high compression ratio above the given value, the CI combustion timing is less likely to fluctuate. Therefore, CI combustion is performed stably, in other words, at a desired timing.

[0017] In reduced cylinder operation, the control unit may actuate an intake valve and an exhaust valve of each of a remainder of the plurality of cylinders.

[0018] This eliminates the need for a valve-stopping mechanism, which incurs high costs. Specifically, since the air / fuel ratio of the mixed gas is lean, the valves of the deactivated cylinder do not need to be stopped to apply the stoichiometric air-fuel ratio, thus reducing pumping loss.

[0019] When a temperature of intake air into the combustion chamber is below a given value, the control unit may perform all-cylinder operation instead of reduced-cylinder operation.

[0020] In this way, instability of the SI combustion at the low temperature of the intake air is eliminated and therefore the combustion timing in the subsequent CI combustion is prevented from being delayed.

[0021] When a temperature of engine cooling water is below a given value, the control unit can perform operation with all cylinders instead of operation with reduced cylinders.

[0022] Thus, in the all-cylinder operation, compared with the reduced-cylinder operation, cooling of unburned mixed gas (end gas) near or in the vicinity of a wall surface of the combustion chamber is promptly canceled, preventing the delay of the combustion start timing of the CI combustion.

[0023] The engine may further include a swirl control valve configured to generate a swirl flow in each combustion chamber. The control unit may control the swirl control valve to generate the swirl flow in the combustion chamber during reduced-cylinder operation.

[0024] By generating or creating the eddy current, the SI combustion is stabilized and the CI combustion is adjusted accordingly. Furthermore, torque changes between combustion cycles are reduced.

[0025] In this case, a vortex ratio of the eddy current may be about 2 or more.

[0026] In this way, as will be described later, the eddy current generated in the combustion chamber is amplified and the SI combustion is carried out stably.

[0027] According to another aspect of the present disclosure, there is provided a compression ignition engine including a plurality of cylinders each formed with a combustion chamber, a spark plug disposed in each combustion chamber, a fuel injection valve disposed to face each combustion chamber, an air-fuel ratio control valve configured to control an air-fuel ratio of a mixture gas within the combustion chamber, and a control unit connected to the spark plug, the fuel injection valve, and the air-fuel ratio control valve and configured to output a control signal to the spark plug, the fuel injection valve, and the air-fuel ratio control valve, respectively. The control unitControl unit includes a processor configured to execute or comprise an engine load determining module to determine whether an engine load is above a given load, an all-cylinder operation control module to, when the engine load is determined to be above the given load, control the fuel injector to inject fuel to all of the plurality of cylinders, and a reduced-cylinder operation control module to, when the engine load is determined to be below the given load, control the fuel injector to interrupt the injection of fuel to one or some of the plurality of cylinders. The reduced-cylinder operation control module includes aAn air-fuel ratio control submodule configured to control the air-fuel ratio control valve to make or set the air-fuel ratio greater than the stoichiometric air-fuel ratio, and a spark plug control submodule configured to output a control signal to the spark plug to perform ignition at a given ignition timing so that the mixture gas starts combustion by flame propagation and then unburned mixture gas self-ignites, wherein the given ignition timing is stored in a memory.

[0028] When the engine is operating at a load below the given load, the control unit switches an operating mode of the engine between a first combustion mode in which SI combustion and CI combustion are performed in this order at an air-fuel ratio substantially equal to or equivalent to a stoichiometric air-fuel ratio, and a second combustion mode in which SI combustion and CI combustion are performed in this order at an air-fuel ratio greater than the stoichiometric air-fuel ratio. When the engine is operating in the first combustion mode, reduced-cylinder operation is not performed.

[0029] According to a further aspect, a method is provided for controlling a compression ignition engine having a plurality of cylinders, comprising the steps of: when the engine is operating at a load above a given load, performing all-cylinder operation by supplying fuel to all of the plurality of cylinders, and when the engine is operating at a load below the given load, performing reduced cylinder operation by supplying fuel to one or some of the plurality of cylinders, and in the reduced-cylinder operation, injecting the fuel to one or some of the plurality of cylinders to produce a mixture gas, and causing the engine, at an air-fuel ratio greater than a stoichiometric air-fuel ratio and at a large compression ratio above a given value, to start spark-ignition combustion in which the mixture gas is ignited by a spark plug to burn by flame propagation, and then to perform compression-ignition combustion in which unburned mixture gas is ignited by self-ignition, when the engine is operating at a load below the given load, switching an operating mode of the engine between a first combustion mode in which the SI combustion and the CI combustion are carried out in this order at an air-fuel ratio substantially equal to a stoichiometric air-fuel ratio, and a second combustion mode in which the SI combustion and the CI combustion are carried out in this order at an air-fuel ratio greater than the stoichiometric air-fuel ratio, wherein when the engine is operating in the first combustion mode, reduced cylinder operation is not performed.

[0030] According to yet another aspect, a control system for a compression ignition engine is provided, which includes a control unit configured to perform the steps of the above-mentioned method.

[0031] According to a further aspect, a computer program product is provided which computer readable instructions which, when loaded onto and executed on a suitable system, can perform the steps of the above-mentioned method. BRIEF DESCRIPTION OF THE DRAWINGS Fig. 1 is a system diagram of an engine and its auxiliaries according to an embodiment of the present disclosure. Fig. 2 shows a plan view (upper part) showing an example of a structure of a combustion chamber of the engine and a cross-sectional view (lower part) taken along a line II-II according to the embodiment of the present disclosure. Fig. 3 is a plan view illustrating an example of structures of the combustion chamber and an intake system of the engine according to the embodiment of the present disclosure. Fig. 4 is a block diagram illustrating an example of a configuration of a control device for the engine according to the embodiment of the present disclosure. Fig. 5 shows a diagram (upper diagram) illustrating operating regions of the engine and an excess air ratio in each operating region, and a diagram (lower diagram) illustrating the operating regions of the engine and an opening of a swirl valve in each operating region according to the embodiment of the present disclosure. Fig. 6 shows timing charts illustrating a fuel injection timing, an ignition timing, and a combustion waveform in each operating region of the engine according to the embodiment of the present disclosure. Fig. 7 is a diagram illustrating a relationship between a first injection timing and a load (average specific effective pressure) in a position 2 of the engine operating region (operating region (1)-2) and a position 3 including a reduced-cylinder operation (operating region (1)-1) according to the embodiment of the present disclosure. Fig. 8 is a diagram illustrating a relationship between a second injection timing and the load in the operating range of the engine at position 2 and the operating range of the engine at position 3 including the reduced cylinder operation according to the embodiment of the present disclosure. Fig. 9 is a diagram illustrating a relationship between a third injection timing and the load in the operating range of the engine at position 2 and the operating range of the engine at position 3 including the reduced cylinder operation according to the embodiment of the present disclosure. Fig. 10 is a diagram illustrating a relationship between a fuel amount of a first injection for one combustion cycle and the load in the operating range of the engine at stage 2 and the reduced-cylinder operation at stage 3 according to the embodiment of the present disclosure. Fig. 11 is a diagram illustrating a relationship between a fuel amount of a second injection for one combustion cycle and the load in the operating range of the engine at stage 2 and the reduced-cylinder operation at stage 3 according to the embodiment of the present disclosure. Fig. 12 is a diagram illustrating a relationship between a fuel amount of a third injection for one combustion cycle and the load in the operating range of the engine at stage 2 and the reduced-cylinder operation at stage 3 according to the embodiment of the present disclosure. Fig. 13 is a diagram illustrating a relationship between a total fuel amount for one combustion cycle and the load in Stage 2 of the engine operating range and Stage 3 including reduced cylinder operation according to the embodiment of the present disclosure. Fig. 14 is a diagram illustrating a relationship between an exhaust valve closing timing and the load in the position 2 of the engine operating range and the position 3 including the reduced-cylinder operation according to the embodiment of the present disclosure. Fig. 15 is a diagram illustrating a relationship between an intake valve opening timing and the load in the position 2 of the engine operating range and the position 3 including the reduced-cylinder operation according to the embodiment of the present disclosure. Fig. 16 is a diagram illustrating a relationship between an additional amount of external EGR and the load in the position 2 of the operating range of the engine and the position 3 including the reduced-cylinder operation according to the embodiment of the present disclosure. Fig. 17 is a flowchart illustrating a control process of a cylinder deactivation operation of the engine according to the embodiment of the present disclosure. DETAILED DESCRIPTION OF THE DISCLOSURE (Embodiment)

[0032] An embodiment of the present invention will be described in detail with reference to the accompanying drawings.

[0033] An embodiment of a control system for a premixed compression ignition engine will be described in detail below with reference to the accompanying drawings. The following description provides an example of the control system for the engine. Fig. Figure 1 is a diagram illustrating a configuration or structure of an engine. Fig. 2 is a view illustrating a structure of a combustion chamber, in which the upper part is a plan view of the combustion chamber and the lower part is a cross-sectional view taken along a line II-II. Fig. Figure 3 is a plan view illustrating structures of the combustion chamber and an intake system. It should be noted that in Fig. 1 an inlet side is or is located on the left side and an outlet side is or is located on the right side of the drawing sheet. Furthermore, Fig. 2 and Fig. 3 the inlet side is on the right side and the outlet side is on the left side of the drawing sheets. Fig. 4 is a block diagram illustrating a configuration of a control device for the motor.

[0034] An engine 1 is a four-stroke engine driven by a combustion chamber 17 that repeats an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke. The engine 1 is mounted on a four-wheeled motor vehicle. The motor vehicle travels by operation of the engine 1. The fuel of the engine 1 is gasoline in this embodiment. The gasoline may contain bioethanol, etc. The fuel of the engine 1 may be any type of fuel as long as it is a liquid fuel containing at least gasoline. (Engine configuration)

[0035] The engine 1 includes a cylinder block 12 and a cylinder head 13, which is arranged and fixed on the cylinder block 12. The cylinder block 12 is formed therein with a plurality of cylinders 11. In Fig. 1 and Fig. 3 illustrates only one cylinder 11. Engine 1 is a multi-cylinder engine.

[0036] A piston 3 is reciprocally mounted in each cylinder 11. The piston 3 is coupled to a crankshaft 15 via a connecting rod 14. The piston 3 defines the combustion chamber 17 together with the cylinder 11 and the cylinder head 13. Note that the definition of a "combustion chamber" is not limited to a space formed when the piston 3 is at a top dead center of a compression stroke (CTDC), but may be broad. That is, "combustion chamber" may mean any space formed by the piston 3, the cylinder 11, and the cylinder head 13, regardless of the position of the piston 3.

[0037] An upper surface of the piston 3 is flat. The piston 3 is recessed in its upper surface to form a cavity 31. The cavity 31 is arranged opposite an injector 6, which will be described later.

[0038] The cavity 31 has a convex cross section 311. The convex cross section 311 is slightly offset from a central axis X1 of the cylinder 11 to an exhaust side. The convex cross section 311 has a substantially conical shape. The convex cross section 311 extends substantially upward in an axis X2 parallel to the central axis X1 of the cylinder 11 from the bottom of the cavity 31. An upper end of the convex cross section 311 is located substantially at the same height as that of an upper surface of a peripheral edge portion of the cavity 31.

[0039] A circumferential side surface of the cavity 31 extends from a bottom surface of the cavity 31 to an opening surface of the cavity 31 inclined relative to the axis X2. An inner diameter of the cavity 31 progressively increases from the bottom surface of the cavity 31 to the opening surface of the cavity 31.

[0040] The cavity 31 has a bottom cross-section 313. An intake-side area or an intake-side surface of the bottom cross-section 313 is opposite a spark plug 25, which will be described later. As shown in the upper part of Fig. 2, the floor cross-section 313 has a given horizontal size.

[0041] Furthermore, as described in the lower part of Fig. As illustrated in FIG. 2, a lower surface of the cylinder head 13, i.e., a ceiling surface of the combustion chamber 17, is formed by an intake-side inclined surface 1311 and an exhaust-side inclined surface 1312. The inclined surface 1311 slopes substantially upward toward the axis X2 from the intake side. On the other hand, the inclined surface 1312 slopes substantially upward toward the axis X2 from the exhaust side. The ceiling surface of the combustion chamber 17 has a so-called pent-roof shape.

[0042] It should be noted that the shape of the combustion chamber 17 is not limited to that shown in Fig. 2. The shapes of the cavity 31, the upper surface of the piston 3, the ceiling surface of the combustion chamber 17, etc., can be suitably changed.

[0043] Furthermore, the cavity 31 may have a symmetrical shape relative to the central axis X1 of the cylinder 11. The inclined surfaces 1311 and 1312 may have a symmetrical shape relative to the central axis X1 of the cylinder 11.

[0044] The geometric compression ratio of the engine 1 is set to be between about 13:1 and about 20:1. As will be described later, within some of the operating ranges of the engine 1, the engine 1 performs SPCCI combustion in which SI combustion and CI combustion are combined. In the SPCCI combustion, CI combustion is performed by utilizing heat generated by SI combustion and pressure rise. In this engine 1, it is not necessary to raise the temperature of the combustion chamber 17 when the piston 3 reaches CTDC for the mixed gas to self-ignite (i.e., compression end temperature). That is, although the engine 1 performs CI combustion, its geometric compression ratio is set to be relatively small.Lowering the geometric compression ratio is beneficial in reducing cooling loss and mechanical loss. For example, the geometric compression ratio of engine 1 can be set to approximately 14:1 to approximately 17:1 in regular specifications (the fuel octane number is approximately 91) and to approximately 15:1 to approximately 18:1 in high-octane specifications (the fuel octane number is approximately 96).

[0045] The cylinder head 13 is formed with an intake opening or an intake port 18 for each cylinder 11. As shown in Fig. 3, the intake port 18 includes two intake ports: a first intake port 181 and a second intake port 182. The first intake port 181 and the second intake port 182 are arranged in axial directions of the crankshaft 15, i.e., forward and backward directions of the engine 1. The intake port 18 communicates with the combustion chamber 17. Although not illustrated in detail, the intake port 18 is a so-called tumble port. That is, the intake port 18 has a shape such that a tumble flow is formed in the combustion chamber 17.

[0046] An intake valve 21 is arranged in the intake port 18. The intake valve 21 opens and closes the intake port 18 to and from the combustion chamber 17. The intake valve 21 is opened and closed by a valve actuating mechanism at a given time. This valve actuating mechanism may be a variable valve mechanism that makes a valve timing and / or a valve lift variable. In this configuration example, as shown in Fig. 1 and Fig. As illustrated in FIG. 4, the variable valve mechanism includes an electrically operated intake S-VT (sequential valve timing) 23. The electrically operated intake S-VT 23 is continuously variable with a rotational phase of an intake camshaft within a given angular range. Thus, the opening and closing timing of the intake valve 21 continuously changes. Note that the actuating mechanism of the intake valve 21 may include a hydraulically operated S-VT instead of the electrically operated S-VT.

[0047] The cylinder head 13 is formed with an exhaust opening or an exhaust port 19 for each cylinder 11. As shown in Fig. 3, the exhaust port 19 also includes two exhaust ports of a first exhaust port 191 and a second exhaust port 192. The first exhaust port 191 and the second exhaust port 192 are arranged in the forward and backward directions of the engine 1. The exhaust port 19 communicates with the combustion chamber 17. An exhaust valve 22 is arranged in the exhaust port 19. The exhaust valve 22 opens and closes the exhaust port 19 to and from the combustion chamber 17. The exhaust valve 22 is opened and closed by a valve operating mechanism at a given time. This valve operating mechanism may be a variable valve mechanism that makes a valve timing and / or a valve lift variable. In this configuration example, as shown in Fig. 1 and Fig. As illustrated in FIG. 4, the variable valve mechanism includes an electrically actuated exhaust S-VT 24. The electrically actuated exhaust S-VT 24 continuously varies with a rotational phase of an exhaust camshaft within a given angular range. Thus, the opening and closing timing of the exhaust valve 22 continuously changes. Note that the exhaust valve actuating mechanism may include a hydraulically actuated S-VT instead of the electrically actuated S-VT.

[0048] The engine 1 adjusts the length of an overlap period between an opening timing of the intake valve 21 and a closing timing of the exhaust valve 22 by the electrically operated intake S-VT 23 and the electrically operated exhaust S-VT 24. Thus, hot burned gas is trapped within the combustion chamber 17. That is, internal EGR (Exhaust Gas Recirculation) gas is introduced into the combustion chamber 17. Further, by adjusting the length of the overlap period, residual gas in the combustion chamber 17 is purged.

[0049] The injector 6 is fixed to the cylinder head 13 for each cylinder 11. The injector 6 injects the fuel directly into the combustion chamber 17. The injector 6 is arranged in a valley portion of the pent roof where the inclined surface 1311 on the intake side and the inclined surface 1312 on the exhaust side intersect each other to be directed into the combustion chamber 17. As shown in Fig. 2, the injector 6 is arranged such that its injection axis is substantially parallel to the central axis X1 of the cylinder. The injection axis of the injector 6 is in line with the axis X2 and coincides with the position of the convex portion 311 of the cavity 31. The injector 6 is oriented or directed toward the cavity 31. It should be noted that the injection axis of the injector 6 may coincide with the central axis X1 of the cylinder 11. Also in this case, it is desirable that the injection axis of the injector 6 coincide with the position of the convex portion 311 of the cavity 31.

[0050] Although not illustrated in detail, the injector 6 is constructed by a multi-port fuel injection valve having a plurality of nozzle openings. As illustrated by a plurality of areas indicated by two-dot chain lines in Fig. 2, the injector 6 injects the fuel such that the fuel spray spreads obliquely downward from the radial center of a ceiling portion of the combustion chamber 17 while radially expanding. In this configuration example, the injector 6 has ten nozzle ports, and the nozzle ports are arranged at a uniform angular interval in the circumferential direction. It should be noted that the number of nozzle ports may be eight. As shown in the upper part of Fig. 2, the axes of the nozzle openings do not circumferentially overlap with the spark plug 25, which will be described later. That is, the spark plug 25 is sandwiched between the axes of two adjacent nozzle openings. This prevents the fuel spray injected from the injector 6 from directly impinging on the spark plug 25 and wetting an electrode.

[0051] A fuel supply system 61 is connected to the injector 6. The fuel supply system 61 includes a fuel tank 63 configured to store the fuel, and a fuel supply path 62 connecting the fuel tank 63 to the injector 6. A fuel pump 65 and a common rail 64 are provided in the fuel supply path 62. The fuel pump 65 pumps the fuel to the common rail 64. In this embodiment, the fuel pump 65 is a motor pump and is arranged inside the fuel tank 63. Furthermore, the fuel pump 65 is connected to a fuel pump control device 651. The common rail 64 stores the fuel pumped by the fuel pump 65 at a high fuel pressure.The common rail 64 is provided with a high-pressure fuel pressure sensor SW16 and a fuel temperature sensor SW161. When the injector 6 opens, the fuel stored in the common rail 64 is injected into the combustion chamber 17 from each nozzle opening of the injector 6. The fuel supply system 61 is capable of supplying fuel at a high pressure of about 30 MPa or higher to the injector 6. A maximum fuel pressure of the fuel supply system 61 may be set to, for example, about 120 MPa. The pressure of the fuel supplied to the injector 6 can be changed according to an operating state of the engine 1. Note that the structure of the fuel supply system 61 is not limited to the above structure.

[0052] Furthermore, a high-pressure fuel pressure pump 641 and a low-pressure fuel sensor SW20 located upstream thereof are arranged in the fuel supply path 62 between the common rail 64 and the fuel pump 65. The high-pressure fuel pressure pump 641 is provided with a fuel temperature sensor SW21.

[0053] The spark plug 25 is fixed to the cylinder head 13 for each cylinder 11. The spark plug 25 forcibly ignites the mixed gas in the combustion chamber 17. In this configuration or embodiment, as shown in Fig. 2, the spark plug 25 is arranged on an intake side of the cylinder 11 relative to the central axis X1 of the cylinder 11. The spark plug 25 is arranged adjacent to or subsequent to the injector 6. The spark plug 25 is arranged between the two intake ports 18. The spark plug 25 is fixed to the cylinder head 13 so as to extend substantially downward toward the center of the combustion chamber 17 in an inclined posture relative to up-down directions of the cylinder head 13. The electrode of the spark plug 25 is arranged near the ceiling surface of the combustion chamber 17 so as to be oriented toward the inside of the combustion chamber 17.

[0054] The cylinder head 13 is provided on the other side (ie, the exhaust side) of the spark plug 25 relative to the central axis X1 of each cylinder 11 with a pressure sensor SW6 which detects a pressure inside the combustion chamber 17.

[0055] An intake passage 40 is connected to one side of the engine 1. The intake passage 40 communicates with the intake ports 18 of the cylinders 11. The intake passage 40 is a passage through which gas to be introduced into the combustion chamber 17 flows. An air cleaner 41 that filters fresh air is disposed in an upstream end portion of the intake passage 40. A surge tank (not illustrated) is disposed near a downstream end of the intake passage 40. A portion of the intake passage 40 downstream of the surge tank forms independent passages that branch for the respective cylinders 11. Downstream ends of the independent passages communicate with the intake ports 18 of the cylinders 11, respectively.

[0056] A throttle valve 43 (an example of an air-fuel ratio control valve) is disposed in the intake passage 40 between the air cleaner 41 and the surge tank. The throttle valve 43 adjusts the intake amount of fresh air into the combustion chamber 17 by adjusting its opening.

[0057] A booster 44 is arranged in the intake passage 40 downstream of the throttle valve 43. The booster 44 compresses the gas to be introduced into the combustion chamber 17. In this embodiment, the booster 44 is a supercharger driven by the engine 1. The booster 44 can be, for example, a Roots type. The booster 44 can have any structure, for example, a Lisholm type, a vane type, or a centrifugal type.

[0058] An electromagnetic clutch 45 is interposed between the booster 44 and the engine 1. The electromagnetic clutch 45 controls, for example, the flow of driving force between the booster 44 and the engine 1, transmitting the driving force from the engine 1 to the booster 44, or interrupting the transmission of driving force therebetween. As described later, the on / off of the booster 44 is switched by an ECU (Engine Control Unit) 10, which switches the connection / disconnection of the electromagnetic clutch 45. In this engine 1, the booster 44 can switch between compressing the gas to be introduced into the combustion chamber 17 and not compressing it.

[0059] An intercooler 46 is disposed in the inlet passage 40 downstream of the booster 44. The intercooler 46 cools the gas that is compressed in the booster 44. The intercooler 46 may, for example, be of a water-cooled type.

[0060] A bypass passage 47 is connected to the inlet passage 40. The bypass passage 47 connects a portion of the inlet passage 40 upstream of the booster 44 with a portion of the inlet passage 40 downstream of the intercooler 46 to bypass the booster 44 and the intercooler 46. An air bypass valve 48 is disposed in the bypass passage 47. The air bypass valve 48 adjusts a flow rate of the gas flowing through the bypass passage 47.

[0061] When the booster 44 is off (i.e., when the electromagnetic clutch 45 is disengaged), the air bypass valve 48 is substantially fully open. Thus, the gas flowing through the intake passage 40 bypasses the booster 44 and is introduced into the combustion chamber 17 of the engine 1. The engine 1 operates in a non-supercharged state, i.e., in a naturally aspirated state.

[0062] When the booster 44 is turned on (i.e., when the electromagnetic clutch 45 is engaged), the gas passing through the booster 44 partially flows back upstream of the booster 44 through the bypass passage 47. By controlling an opening of the air bypass valve 48, a backflow amount is adjusted, which results in adjusting a boost pressure of the gas introduced into the combustion chamber 17. In this embodiment, a boosting system 49 consists of the booster 44, the bypass passage 47, and the air bypass valve 48.

[0063] The engine 1 has a vortex generating part which generates a vortex flow in the combustion chamber 17. In one example, as shown in Fig. 3, the swirl generating part is a swirl control valve (SCV) 56 fixed to the intake passage 40. The SCV 56 is arranged in a passage. The passage is one of a primary passage 401 and a secondary passage 402, which communicate with the first intake port 181 and the second intake port 182, respectively. The SCV 56 is an opening regulating valve capable of adjusting an opening of a cross-sectional area of ​​the secondary passage 402. When the opening of the SCV 56 is small, the flow rate of intake air into the combustion chamber 17 from the first intake port 181 relatively increases, while the flow rate of intake air into the combustion chamber 17 from the second intake port 182 is relatively reduced. Thus, the swirl flow in the combustion chamber 17 becomes strong.When the opening of the SCV 56 is large, the flow rates of intake air into the combustion chamber 17 from the first intake port 181 and the second intake port 182 become substantially equal, and thus the swirl flow in the combustion chamber 17 becomes weak. When the SCV 56 is substantially fully open, swirl flow does not occur. Note that the swirl flow is counterclockwise in . Fig. 3 circulates, as indicated by the arrows (see also the white arrows with outer contour in Fig. 2).

[0064] Note that, alternatively / in addition to fixing the SCV 56 to the intake passage 40, the swirl generating part may adopt a structure in which the open periods of the two intake valves 21 are varied to introduce the intake air into the combustion chamber 17 from only one of the intake valves 21. By opening only one of the two intake valves 21, the intake air is unevenly introduced into the combustion chamber 17, thus generating the swirl flow in the combustion chamber 17. Alternatively, the shapes of the intake ports 18 may be designed so that the swirl generating part generates the swirl flow in the combustion chamber 17.

[0065] An exhaust passage 50 is connected to a side of the engine 1 opposite to the intake passage 40. The exhaust passage 50 communicates with the exhaust ports 19 of the cylinder 11. The exhaust passage 50 is a passage through which the exhaust gas discharged from the combustion chamber 17 flows. Although not illustrated in detail, an upstream part of the exhaust passage 50 constitutes independent passages branched to the respective cylinders 11. Upstream ends of the independent passages are connected to the exhaust ports 19 of the cylinders 11, respectively.

[0066] An exhaust gas purification system including a plurality of catalytic converters is disposed in the exhaust passage 50. Although not illustrated, the upstream catalytic converter is disposed in an engine compartment and includes a three-way catalyst 511 and a GPF (Gasoline Particulate Filter) 512. The downstream catalytic converter is disposed outside the engine compartment and includes a three-way catalyst 513. Note that the exhaust gas purification system is not limited to having the illustrated structure.

[0067] An EGR passage 52, which constitutes an external EGR system, is connected between the intake passage 40 and the exhaust passage 50. The EGR passage 52 is a passage for recirculating a portion of the combusted gas to the intake passage 40. An upstream end of the EGR passage 52 is connected to the exhaust passage 50 between the upstream catalytic converter and the downstream catalytic converter. A downstream end of the EGR passage 52 is connected to the intake passage 40 upstream of the booster 44.

[0068] A water-cooling type EGR cooler 53 is disposed in the EGR passage 52. The EGR cooler 53 cools the burned gas. An EGR valve 54 (another example of the air-fuel ratio control valve) is also disposed in the EGR passage 52. The EGR valve 54 adjusts the flow rate of the burned gas in the EGR passage 52. By adjusting an opening of the EGR valve 54, the recirculation amount of the cooled burned gas (i.e., external EGR gas) is adjusted.

[0069] In this embodiment, an EGR system 55 includes an external EGR system including the EGR passage 52 and the EGR valve 54, and an internal EGR system including the electrically operated intake S-VT 23 and the electrically operated exhaust S-VT 24 described above.

[0070] A control system 20 for the compression ignition engine includes the ECU 10, which is configured to operate the engine 1. As shown in Fig. 4, the ECU 10 is a control unit based on a well-known microcomputer and includes a processor (e.g., a central processing unit (CPU)) 101 configured to execute program(s) / instructions, a memory 102 consisting of RAM(s) (random access memory) and ROM(s) (read-only memory) and configured to store the program(s) / instructions and data, an input / output bus 103 configured to input and output electrical signals, an engine load determining module 104, a full-cylinder operation control sub-module 105, and a reduced-cylinder operation control module 106. The control module 106 for operation with reduced cylinders or the module 106 controlling or regulating the operation with reduced cylinders comprises a control orAn air-fuel ratio control submodule 107 and a spark plug control submodule 108 are included. Processor 101 is configured to execute these modules and submodules to perform their respective functions. These modules and submodules are stored in memory 102 as programs. ECU 10 is an example of a "control unit."

[0071] As this is Fig. 1 and Fig. As illustrated in Figure 4, various sensors such as SW1 to SW17, SW20 to SW24, SW31, SW51, SW101, SW102, and SW161 are connected to the ECU 10. These sensors output detection signals to the ECU 10. The sensors include, for example, the following sensors.

[0072] That is, the sensors include an air flow sensor SW1, which is arranged in the intake passage 40 downstream of the air cleaner 41 and configured to detect the flow rate of fresh air in the intake passage 40, a first intake air temperature sensor SW2, which is also arranged in the intake passage 40 downstream of the air cleaner 41 and configured to detect a temperature of the fresh air, an intake air pressure sensor SW3, which is arranged in the intake passage 40 downstream of the connecting position with the EGR passage 52 and upstream of the booster or compressor 44 and configured to detect a pressure of the gas flowing into the booster 44, a second intake air temperature sensor SW31,which is arranged in the intake passage 40 downstream of the connecting position with the EGR passage 52 and upstream of the booster 44 and is configured to detect a temperature of the gas, a third intake air temperature sensor SW4 which is arranged in the intake passage 40 downstream of the booster 44 and upstream of the intercooler 46 and is configured to detect a temperature of the gas flowing out of the booster 44, a booster pressure sensor SW5 which is attached to the surge tank which is arranged downstream of the intercooler 46 and is configured to detect a pressure of the gas at a position downstream of the booster 44, a fourth intake air temperature sensor SW51 which is attached to the surge tank which is arranged downstream of the intercooler 46 and is configured,to detect a temperature of the gas, the pressure sensors SW6, which are attached to the cylinder head 13 corresponding to the cylinders 11 and are configured to detect a pressure in the combustion chambers 17, respectively; an exhaust or exhaust gas temperature sensor SW7, which is arranged in the exhaust passage 50 and is configured to detect a temperature of the exhaust gas discharged from the combustion chamber 17; a linear O2 sensor SW8, which is arranged in the exhaust passage 50 upstream of the upstream catalytic converter and is configured to detect an oxygen concentration within the exhaust gas; a lambda O2 sensor SW9, which is arranged in the upstream catalytic converter downstream of the three-way catalyst 511 and is configured to detect an oxygen concentration within the exhaust gas after it has passed through the three-way catalyst 511;a first water temperature sensor SW10 (see the upper right part of , Fig. 1) which is attached to the cylinder head 13 and configured to detect a temperature of cooling water, a second water temperature sensor SW101 which is attached near a discharge port of the cylinder head 13 leading toward a main cooler and configured to detect a temperature of the cooling water, a third water temperature sensor SW102 which is attached near a discharge port leading toward a water pump W / P of the main cooler and configured to detect a temperature of cooling water, a crank angle sensor SW11 which is attached to the engine 1 and configured to detect a rotation angle of the crankshaft 15, an accelerator opening sensor SW12 which is attached to an accelerator pedal mechanism and configured to detect an accelerator opening according to an operation amount of an accelerator pedal.accelerator pedal, an intake cam angle sensor SW13 which is fixed to the engine 1 and configured to detect a rotation angle of the intake camshaft, an exhaust cam angle sensor SW14 which is fixed to the engine 1 and configured to detect a rotation angle of the exhaust camshaft, an EGR pressure difference sensor SW15 which is arranged in the EGR passage 52 and configured to detect a difference in pressure between positions upstream and downstream of the EGR valve 54, a fuel pressure sensor SW16 which is fixed to the common rail 64 of the fuel supply system 61 and configured to detect a pressure of the fuel to be supplied to or discharged from the injector 6.to be supplied, a fuel temperature sensor SW161 which is attached to the common rail 64 of the fuel supply system 61 and configured to detect a temperature of the fuel to be supplied to the injector 6, and a throttle opening sensor SW17 which is attached to a drive motor of the throttle valve 43 and configured to detect an opening of the throttle valve 43.

[0073] In addition, the sensors include the low-pressure fuel pressure sensor SW20 installed on the fuel supply path 62 between the high-pressure fuel pump 641 and the fuel pump 65, the fuel temperature sensor SW21 installed on the high-pressure fuel pump 641, a GPF pressure sensor SW22 installed on the GPF 512, a hydraulic pressure sensor SW23 installed on the cylinder block 12, and an oil level sensor SW24 installed on the bottom of the oil pan.

[0074] Based on these detection signals, the ECU 10 determines the operating state of the engine 1 and calculates control variables for the various devices. The ECU 10 outputs control signals related to the calculated control variables to the injector 6, the spark plug 25, the electrically actuated intake S-VT 23, the electrically actuated exhaust S-VT 24, the fuel supply system 61, the throttle valve 43, the EGR valve 54, the electromagnetic clutch 45 of the booster 44, the air bypass valve 48, and the SCV 56. For example, the ECU 10 adjusts the compressor pressure by controlling an opening of the air bypass valve 48 based on a pressure difference between the upstream side and the downstream side of the booster 44 obtained from the detection signals of the intake air pressure sensor SW3 and the boosting pressure sensor SW5.Furthermore, the ECU 10 adjusts the amount of external EGR gas introduced into the combustion chamber 17 by controlling the opening of the EGR valve 54 based on a pressure difference between the upstream and downstream sides of the EGR valve 54, which is obtained from the detection signal of the EGR pressure difference sensor SW15. Details of the control of the engine 1 by the ECU 10 will be described later. (Engine operating range)

[0075] Fig. 5 illustrates operating range maps 501 and 502 of the engine 1. The operating range maps 501 and 502 of the engine 1 are divided into five ranges with respect to the engine load and the engine speed. For example, the five ranges include: a low load range (1)-1, which includes idling operation and extends into low and medium engine speed ranges; a medium load range (1)-2, in which the engine load is higher than in the low load range and which extends into low and medium engine speed ranges; a high load and medium speed range (2), in which the engine load is higher than in the medium load range (1)-2 and which is located in a high load range and includes a full engine load; a high load and low speed range (3).speed, which is located in the high load region and in which the engine speed is lower than in the high load medium speed region (2); and a high speed region (4) in which the engine speed is higher than in the light load region (1)-1, the medium load region (1)-2, the high load medium speed region (2), and the high load low speed region (3). Here, the low speed region, the medium speed region, and the high speed region can be defined by substantially equally dividing in the direction of the engine speed the entire operating region of the engine 1 into three regions of the low speed region, the medium speed region, and the high speed region. In the example of . Fig. 5 If the engine speed is lower than a speed N1, it is defined as low, if the engine speed is higher than a speed N2, it is defined as high, and if the engine speed is between the engine speeds N1 and N2, it is defined as medium. For example, the speed N1 can be approximately 1,200 rpm (min -1 ) and the speed N2 can be about 4,000 rpm (min -1 ). The area (2) of high load and medium speed may be an area where the combustion pressure exceeds approximately 900 kPa. In Fig. 5, the operating range maps 501 and 502 of the engine 1 are illustrated separately for the purpose of easier understanding. Map 501 illustrates a state of the mixed gas and a combustion mode within each range, and a driving range and a non-driving range of the booster 44. Map 502 illustrates the opening of the SCV 56 within each range. Note that the two-dot dashed line in Fig. 5 shows the road load line of engine 1.

[0076] Mainly for improving fuel efficiency and exhaust performance, the engine 1 performs combustion by compression self-ignition (ie, SPCCI combustion) within the (1)-1 low load region, the (1)-2 medium load region, and the (2) high load and medium speed region.

[0077] In SPCCI combustion, the spark plug 25 forcibly ignites the mixed gas inside the combustion chamber 17 to undergo SI combustion by flame propagation, and the heat generated by this combustion increases the temperature inside the combustion chamber 17 and the pressure inside the combustion chamber 17 increases by flame propagation, resulting in CI combustion of unburned mixed gas by self-ignition.

[0078] It is possible to reduce the temperature change inside the combustion chamber 17 before compression starts by adjusting the heat generation amount in the SI combustion. Therefore, even if the temperature inside the combustion chamber 17 varies before compression starts, for example, by controlling the ignition timing to adjust the start timing of the SI combustion, the mixed gas itself ignites at a target timing.

[0079] In SPCCI combustion, heat generation in SI combustion is slower than heat generation in CI combustion. A waveform of a heat generation rate in SPCCI combustion has a relatively gentle slope, as indicated by reference numerals 6014, 6024, 6034, and 6063 in Fig. 6. Moreover, a pressure change (dp / dθ) in the combustion chamber 17 is gentler in the SI combustion than in the CI combustion. In other words, the heat generation rate waveform of the SPCCI combustion is formed to include a first heat generation rate portion, which is formed by the SI combustion and has a relatively gentle increasing slope, and a second heat generation rate portion, which is formed by the CI combustion and has a relatively sharp increasing slope, which are adjacent to each other in this order.

[0080] When the temperature and pressure inside the combustion chamber 17 increase due to SI combustion, the unburned mixed gas ignites itself. In the example of the waveforms 6014, 6024, 6034 and 6063 of the heat generation rate, which are shown in Fig. As illustrated in Figure 6, the slope of the waveform changes from gentle to sharp, or strong, depending on the timing of auto-ignition. That is, the heat generation rate waveform exhibits an inflection point at the time when CI combustion starts.

[0081] After the CI combustion starts, the SI combustion and the CI combustion are performed in parallel. In the CI combustion, since the heat generation is greater than in the SI combustion, the heat generation rate becomes relatively high. It should be noted that since the CI combustion is performed after the CTDC, the piston 3 moves downward by being driven. Therefore, the inclinations of the waveforms 6014, 6024, 6034, and 6063 of the heat generation rate by the CI combustion are prevented from becoming excessively strong. The dp / dθ in the CI combustion also becomes relatively gentle.

[0082] The dp / dθ can be used as an index that expresses combustion noise. Since SPCCI combustion is capable of reducing the dp / dθ as described above, it is possible to prevent combustion noise from becoming excessively loud. Thus, combustion noise is suppressed below an acceptable level.

[0083] SPCCI combustion ends with the completion of CI combustion. CI combustion has a shorter combustion period than SI combustion. SPCCI combustion advances the combustion end timing compared to SI combustion. In other words, SPCCI combustion brings the combustion end timing on or in the expansion stroke closer to the CTDC. SPCCI combustion is advantageous in improving the fuel efficiency of engine 1 compared to SI combustion.

[0084] Furthermore, in the SPCCI combustion, a combustion control device (EGR, VVT, intake air amount control device) is controlled to change the heat generation amount of the SI combustion (the first portion of the heat generation rate) according to the operating state of the engine, so that the start timing of the CI combustion (the second portion of the heat generation rate) reaches a target CI combustion start timing, which is set according to the operating state of the engine.

[0085] Furthermore, the engine 1 performs the SI combustion operation by spark ignition within the other regions, specifically the high load and low speed region (3) and the high speed region (4). The operation of the engine 1 within each region will be described in detail with reference to the fuel injection timing and the ignition timing, which are Fig. 6 are illustrated.

[0086] Fig. 6 shows a fuel injection timing and an ignition timing within each range of the operating range maps 501 and 502 of Fig. 5. Reference numerals 601, 602, 603, 604, 605 and 606 of Fig. 6 correspond to operating states 601, 602, 603, 604, 605 and 606 of Fig. 5. The operating state 606 corresponds to an operating state where an engine speed is high within the high load and medium speed range (2). (Range (1)-1 low load)

[0087] When engine 1 operates within the low-load range (1)-1, engine 1 performs the CI combustion operation as described above. In combustion caused by self-ignition, the timing of self-ignition changes greatly as the temperature inside the combustion chamber varies before compression starts. Therefore, within the low-load range (1)-1, engine 1 performs the SPCCI combustion operation in which SI combustion and CI combustion are combined.

[0088] When the engine 1 operates in the operating state 601 of the low-load region (1)-1, the injector 6 injects the fuel into the combustion chamber 17 by dividing it into a plurality of injections on the compression stroke (see reference numerals 6015 and 6016). The mixed gas is stratified in a central portion and an outer peripheral portion of the combustion chamber 17 by the divided injections of the fuel and a strong swirl flow in the combustion chamber 17.

[0089] After the fuel injection is completed, the spark plug 25 ignites the mixed gas in the central portion of the combustion chamber 17 at a given timing before CTDC (see reference numeral 6013). Since the fuel concentration of the mixed gas in the central portion is relatively high, ignitability improves, and SI combustion stabilizes through flame propagation. By stabilizing SI combustion, CI combustion starts at the appropriate timing (see combustion waveform 6014). The controllability of CI combustion improves in SPCCI combustion. As a result, when the engine 1 operates within the (1)-1 low load range, both the reduction in generation of combustion noise and the improvement in fuel efficiency by shortening the combustion period are achieved.

[0090] Since the engine 1 performs the SPCCI combustion while the mixture gas is lean within the (1)-1 low load region, the (1)-1 low load region can be called “SPCCI λ>1 region”, ie, a second combustion mode. (range (1)-2 medium load)

[0091] Even when the engine 1 operates within the medium load range (1)-2, the engine 1 performs the SPCCI combustion process similar to the low load range (1)-1.

[0092] When the engine 1 operates in the operating state 602 within the medium load range (1)-2, the injector 6 performs fuel injection on the intake stroke (see reference numeral 6025) and fuel injection on the compression stroke (see reference numeral 6026). By performing a first injection 6025 on the intake stroke, the fuel is distributed substantially evenly into the combustion chamber 17. By performing a second injection 6026 on the compression stroke, at a high engine load within the medium load range (1)-2, the temperature inside the combustion chamber 17 is lowered by latent heat of evaporation of the fuel, and thus abnormal combustion such as knocking is prevented. A ratio between the injection amount of the first injection 6025 and the injection amount of the second injection 6026 may be, for example, approximately 95:5.

[0093] When the injector 6 performs the first injection 6025 on the intake stroke and the second injection 6026 on the compression stroke, the mixed gas with the excess air ratio λ of 1.0±0.2 is formed in the combustion chamber 17 as a whole. Since the fuel concentration of the mixed gas is substantially homogeneous, the improvement in fuel efficiency by reducing the loss of unburned fuel and the improvement in exhaust performance by preventing smoke generation are achieved. The excess air ratio λ is preferably about 1.0 to about 1.2.

[0094] Because the spark plug 25 ignites the mixed gas at a given time before the CTDC (see reference numeral 6023), the mixed gas burns through flame propagation. After combustion is initiated through flame propagation, the unburned mixed gas ignites itself at the target time, causing CI combustion (see combustion waveform 6024).

[0095] Within the medium load region (1)-2, since the engine 1 performs the SPCCI combustion by adjusting the mixed gas substantially to a stoichiometric air-fuel ratio, the medium load region (1)-2 can be referred to as an “SPCCI λ=1 region”, ie, a first combustion mode.

[0096] Here is how this is done in map 501 of Fig. 5, the booster 44 is switched off (see “S / C OFF” in Fig. 5) within a part of the low load range (1)-1 and a part of the medium load range (1)-2. In detail, the booster 44 is turned off within a low speed segment of the low load range (1)-1. Within a high speed segment of the low load range (1)-1, the booster 44 is turned on to increase the boost pressure to ensure a required intake charge amount corresponding to the engine speed, which is high. Furthermore, within a low load and low speed segment of the medium load range (1)-2, the booster 44 is turned off. Within a high load segment of the medium load range (1)-2, the booster 44 is turned on to ensure the required intake charge amount corresponding to the engine speed, which is high.Furthermore, within the high speed segment, the booster 44 is switched on to ensure the required intake charge amount according to the engine speed, which is high.

[0097] It should be noted that within the high load and medium speed range (2), the high load and low speed range (3) and the high speed range (4), the booster 44 is switched on throughout the ranges. (Area (2) high load and medium speed)

[0098] Even when the engine 1 operates within the high load and medium speed region (2), the engine 1 performs the SPCCI combustion process similarly to the low load region (1)-1 and the medium load region (1)-2.

[0099] When the engine 1 operates in the operating state 603 at a low speed segment of the high load and medium speed region (2), the injector 6 injects the fuel at the intake stroke (see reference numeral 6035) and injects the fuel at a final stage of the compression stroke (see reference numeral 6036).

[0100] The first-stage injector 6035, which starts on the intake stroke, can start fuel injection in the early half of the intake stroke. The early half of the intake stroke can be defined by equally dividing the intake stroke into two parts, the early half and the late half. For example, the first-stage injector can start fuel injection at approximately 280° CA (crank angle) before TDC (top dead center).

[0101] When the injection of the first-stage injector 6035 is started in the early half of the intake stroke, the fuel spray impinges on an opening edge of the cavity 31, so that a portion of the fuel enters a squish area 171 of the combustion chamber 17, and the rest of the fuel enters the cavity 31. The swirling flow is strong in the outer peripheral portion of the combustion chamber 17 and weak in the central portion. Therefore, the portion of the fuel that has entered the squish area 171 joins the swirling flow, and the rest of the fuel that has entered the cavity 31 joins the inner side of the swirling flow. The fuel that has joined the swirling flow remains in the swirling flow during the intake stroke and the compression stroke, and forms the mixed gas for CI combustion in the outer peripheral portion.Portion of the outer periphery of the combustion chamber 17. The fuel that has entered the inner side of the swirling flow also remains on the inner side of the swirling flow during the intake stroke and the compression stroke and forms the mixed gas for SI combustion in the central portion of the combustion chamber 17.

[0102] The excess air ratio λ of the mixed gas in the central portion where the spark plug 25 is arranged is preferably about 1 or less, and the excess air ratio λ of the mixed gas in the outer peripheral portion is about 1 or less, preferably below 1. The air-fuel ratio (A / F) of the mixed gas in the central portion may be, for example, between about 13 and the stoichiometric air-fuel ratio (14.7:1). The air-fuel ratio of the mixed gas in the central portion may be leaner than the stoichiometric air-fuel ratio. Furthermore, the air-fuel ratio of the mixed gas in the outer peripheral portion may be, for example, between about 11:1 and the stoichiometric air-fuel ratio, preferably between about 11:1 and about 12:1.The air-fuel ratio of the mixed gas in the entire combustion chamber 17 may be between about 12.5:1 and the stoichiometric air-fuel ratio, preferably between about 12.5:1 and about 13:1.

[0103] The second-stage injection 6036, which is performed in the final stage of the compression stroke, can start fuel injection at 10° CA before TDC. By performing the second-stage injection immediately before TDC, the temperature in the combustion chamber 17 is lowered by the latent heat of fuel vaporization. A low-temperature oxidation reaction of the fuel injected by the first-stage injection 6035 proceeds in the compression stroke and transitions to a high-temperature oxidation reaction before TDC. Here, by performing the second-stage injection 6036 immediately before TDC to lower the temperature inside the combustion chamber, the transition from the low-temperature oxidation reaction to the high-temperature oxidation reaction is avoided, and preignition is prevented.It should be noted that the ratio between the injection quantity of the first stage injection 6035 and the injection quantity of the second stage injection 6036 may be, for example, approximately 95:5.

[0104] The spark plug 25 ignites the mixed gas in the central portion of the combustion chamber 17 near the CTDC (see reference numeral 6037). The spark plug 25 ignites, for example, after the CTDC. Since the spark plug 25 is disposed in the central portion of the combustion chamber 17, the ignition of the spark plug 25 causes the mixed gas in the central portion to start SI combustion through flame propagation. The flame of SI combustion propagates in the circumferential direction on a strong swirling current inside the combustion chamber 17. At a given position of the outer peripheral portion of the combustion chamber 17 in the circumferential direction, the unburned mixed gas is compressed and ignited, and CI combustion starts (see combustion waveform 6034).

[0105] In this regard, when the engine 1 operates in the operating state 606 at the high speed side of the high load and medium speed region (2), the injector 6 starts the fuel injection on the intake stroke (see reference numeral 6061).

[0106] The first-stage injection 6061, which starts on the intake stroke, may start fuel injection in the early half of the intake stroke, similar to the first-stage injection 6035 in the operating state 603. For example, the first-stage injection 6061 may start fuel injection at approximately 280° CA before TDC. The first-stage injection 6061 may extend over the intake stroke and end on the compression stroke. By setting the start of injection of the first-stage injection 6061 in the early half of the intake stroke, the mixed gas for CI combustion is formed in the outer peripheral portion of the combustion chamber 17, and the mixed gas for SI combustion is formed in the central portion of the combustion chamber 17. Abnormal combustion does not easily occur due to the high engine speed.-speed, and the second stage injection can be omitted.

[0107] The spark plug 25 ignites the mixed gas in the central portion of the combustion chamber 17 near the CTDC (see reference numeral 6062). The spark plug 25 ignites, for example, after the CTDC. Thus, SPCCI combustion is performed (see combustion waveform 6063).

[0108] Within the high load and medium speed range (2), since the engine 1 performs the SPCCI combustion process while the mixed gas is richer, the high load and medium speed range (2) can be called “SPCCI λ<=1 range”. (Area (3) high load and low speed)

[0109] When the engine speed is low, the time required for the crank angle to change by 1° becomes longer. Within the high load and low speed range (3), when fuel is injected into the combustion chamber 17 in the intake stroke and the early half of the compression stroke, similar to the high load and medium speed range (2), the fuel reaction proceeds excessively and causes preignition. When the engine 1 operates within the high load and low speed range (3), it becomes difficult to perform the SI combustion process.

[0110] Therefore, when the engine 1 operates within the high load and low speed range (3), the engine 1 performs the SI combustion process instead of the SPCCI combustion process.

[0111] When the engine 1 operates in the operating state 604 within the high load and low speed range (3), the injector 6 injects the fuel into the combustion chamber 17 at the timing of the intake stroke and in a delay period from the final stage of the compression stroke to the early stage of the expansion stroke (see reference numerals 6044 and 6045). By injecting the fuel in two injections as described above, the amount of fuel injected in the delay period is reduced. By injecting the fuel in the intake stroke (see reference numeral 6044), the time period of formation of the mixed gas is sufficiently ensured. In addition, by injecting the fuel in the delay period (see reference numeral 6045), the flowThe flow in the combustion chamber 17 is amplified immediately before ignition, which is advantageous in stabilizing SI combustion. This fuel injection mode is particularly effective when the geometric compression ratio of the engine 1 is low.

[0112] After the fuel is injected, the spark plug 25 ignites the mixed gas at a timing close to the CTDC (see reference numeral 6042). For example, the spark plug 25 ignites after the CTDC. The mixed gas causes SI combustion on the expansion stroke. Since SI combustion starts on the expansion stroke, CI combustion does not start (see combustion waveform 6043).

[0113] Within the high load and low speed region (3), since the engine 1 performs the SI combustion process by injecting the fuel in the delay period from the final stage of the compression stroke to the early stage of the expansion stroke, the high load and low speed region (3) can be referred to as a “delayed SI region”. (Area (4) high speed)

[0114] When the engine speed is high, the time required for the crank angle to change by 1° becomes shorter. Therefore, for example, within a high-speed segment of the high-load region, it is difficult to stratify the mixed gas in the combustion chamber 17 by performing the divided injections during the compression stroke. That is, as the engine speed increases, it becomes difficult to perform the SPCCI combustion process described above. Therefore, when the engine 1 operates within the high-speed region (4), the engine 1 performs the SI combustion process instead of the SPCCI combustion process. Note that the high-speed region (4) extends across the entire load direction from a low load to a high load.

[0115] The reference number 605 in Fig. 6 shows an example of the fuel injection timing (refer to reference numeral 6051) and the ignition timing (refer to reference numeral 6052) and a combustion waveform (refer to reference numeral 6053) when the engine 1 operates in a high load segment of the high speed region (4).

[0116] The EGR system 55 introduces EGR gas into the combustion chamber 17 when the operating state of the engine 1 is within the high-speed range (4). The engine 1 reduces the amount of EGR gas as the engine load increases. At full load, the EGR gas can be set to approximately zero.

[0117] When operating within the high-speed range (4), the engine 1 substantially fully opens the SCV 56. No swirl flow is generated in the combustion chamber 17, and only the tumble flow is generated. By substantially fully opening the SCV 56, the charging efficiency in the high-speed range (4) is improved and pumping loss is reduced.

[0118] When the engine 1 operates within the high-speed range (4), the air-fuel ratio (A / F) of the mixed gas is substantially at the stoichiometric air-fuel ratio (A / F approximately 14.7:1) throughout the combustion chamber 17. That is, the excess air ratio λ of the mixed gas may be 1.0±0.2. It should be noted that within a high-load segment of the high-speed range (4) including the full load, the excess air ratio λ of the mixed gas may be less than 1.

[0119] When the engine 1 is operating within the high-speed range (4), the injector 6 starts fuel injection on the intake stroke. The injector 6 injects all the fuel for one combustion cycle together. It should be noted that the reference numeral 6051 of Fig. 6 indicates a fuel injection state where the engine load is high and the fuel injection amount is large, and the fuel injection period changes according to the fuel injection amount. By starting fuel injection on the intake stroke, a homogeneous or substantially homogeneous mixed gas is formed in the combustion chamber 17. Furthermore, when the engine speed is high, since the evaporation time of the fuel is ensured as long as possible, the loss of unburned fuel and the generation of soot are reduced.

[0120] After the fuel injection is completed, the spark plug 25 ignites the mixture gas at an appropriate timing before the CTDC (see reference numeral 6052).

[0121] Within the high speed region (4), since the engine 1 starts the fuel injection on the intake stroke and performs the SI combustion, the high speed region (4) can be referred to as “intake SI region”. (“Operation with reduced cylinders” in the range (1)-1 low load)

[0122] In this embodiment, as shown in Fig. 5, the reduced-cylinder operation is performed within a lowest load segment of the low-load region (1)-1 (referred to herein as "Layer 3" for convenience). The lowest load segment of Layer 3 may, for example, be a region of approximately 200 kPa and below in terms of a brake mean effective pressure (BMEP). It should be noted that the BMEP does not indicate the engine load itself, but is a value obtained by multiplying the BMEP by an exhaust gas amount proportional to an axial torque.

[0123] Within such an operating range where the BMEP is about 200 kPa or lower, since a throttle loss (pumping loss) when the engine 1 is combusting increases, the reduced-cylinder operation (cylinder deactivation operation) in which operation of some of the plurality of cylinders (e.g., two of four cylinders) is suspended is performed. Here, the throttle valve 43 is also opened for the suspended cylinders. For the operating cylinder, as described above, the compression ratio (geometric compression ratio) of the mixed gas can be set comparatively large to about 13.0:1 or more.

[0124] Since such reduced cylinder operation increases the load of the two suspended cylinders during operation, the throttling loss is reduced.

[0125] It should be noted that the intake and exhaust valves of the suspended cylinders can be in the operating state. With this configuration, since a valve stop mechanism is not required, the manufacturing cost of the engine capable of operating with reduced cylinders is reduced.

[0126] During reduced-cylinder operation, although the booster 44 does not perform compression, the electromagnetic clutch 45 of the booster 44 may remain connected regardless of the engine speed, so that a mechanical load due to connecting / disconnecting the electromagnetic clutch 45 decreases.

[0127] Furthermore, as shown in map 502 of Fig. As illustrated in Figure 5, during reduced-cylinder operation, swirl flow can be generated in the operating cylinders. In this case, the swirl ratio of the swirl flow can be set to approximately 4 or higher. Thus, SI combustion is stably performed at an excess air ratio λ>1, i.e., lean A / F, and the swirl flow within the combustion chamber is enhanced, further stabilizing SI combustion.

[0128] When the intake air temperature into the combustion chamber is below a given value, full-cylinder operation can be performed instead of reduced-cylinder operation. This eliminates the instability of SI combustion at low intake air temperatures, preventing the combustion timing from being delayed in subsequent CI combustion.

[0129] Furthermore, when the cooling water temperature of the engine 1 is below a given value, full-cylinder operation can be performed instead of reduced-cylinder operation. Thus, in full-cylinder operation, compared to reduced-cylinder operation, cooling of the end gas (unburned mixture gas) near or near the wall surface of the combustion chamber is promptly canceled, preventing the delay in the combustion start timing of the CI combustion.

[0130] In addition, as stated in Fig. 5, within the medium load range (1)-2 (referred to as “position 2” for convenience), the SI combustion and the CI combustion are performed in that order at the air-fuel ratio equivalent to the stoichiometric air-fuel ratio (i.e., λ=1).

[0131] Here, the reduced-cylinder operation may or may not be performed even in a case where the current operating state shifts from position 2, where the operation is performed at the air-fuel ratio equivalent to the excess air ratio λ=1, to position 3 according to the operating state of the engine 1. That is, even if the operating state shifts to position 3 while the engine 1 is operating in position 2 equivalent to λ=1, the all-cylinder operation is continued without shifting or switching to the reduced-cylinder operation. In this way, generation of NO x when switching the control system to operation with reduced cylinders.

[0132] Next, relationships between the engine load and the fuel quantity in position 2 and position 3 are described with reference to the drawings.

[0133] Fig. 7 to 9 illustrate an example of a relationship between a first to third injection timing and each load in position 3, which includes operation with reduced cylinders (cylinder deactivation), and position 2. Further illustrate Fig. 10 to 12 show an example of a relationship between the fuel quantity for a combustion cycle and the load for a plurality of fuel injection timings. The engine speed or rpm in each diagram of Fig. For example, 7 to 12 is about 3,500 rpm. This is similar for Fig. 13 to 16, which will be described later.

[0134] For example, as shown in Fig. 7, the first injection timing of the three injection timings may be approximately 88° CA before CTDC (bTDC). At a further lower engine load, where the operating state shifts to position 3, where cylinder deactivation is performed, the first injection timing may be approximately 320° CA before CTDC and advanced to an early stage of the intake stroke. The first injection timing may be retarded to approximately 90° CA before CTDC as the engine load decreases.

[0135] In contrast, as shown in Fig. As illustrated in Figure 10, the fuel amount of a first injection progressively decreases as the engine load decreases in position 3, where the operation is performed with all cylinders, and increases to approximately six times at an engine load at which the operation state switches to cylinder deactivation. The fuel amount of the first injection is reduced as the engine load further decreases.

[0136] As this is Fig. As illustrated in Figure 8, the timing of the second injection of the three injection timings may be approximately 60 to 50° CA before CTDC. At a further lower engine load, where the operating state switches to position 3, where cylinder deactivation is performed, the second injection timing may be approximately 250° CA before CTDC. The second injection timing may be retarded to approximately 60° CA before CTDC as the engine load decreases.

[0137] Meanwhile, or in contrast, as in Fig. As illustrated in Figure 11, the fuel quantity of the second injection progressively decreases within the low-load segment of position 3, where all-cylinder operation is performed, and is further reduced at the engine load at which the operating state switches to cylinder deactivation. The fuel quantity of the second injection is increased as the engine load further decreases.

[0138] As this is Fig. As illustrated in Figure 9, the timing of the third injection of the three injection timings may be approximately 45° CA before CTDC within the low-load segment of Stage 3, where all-cylinder operation is performed. At the further lower engine load, at which the operating state switches to Stage 3, where cylinder deactivation is performed, the third injection timing may be approximately 200° CA before CTDC. The third injection timing may be retarded to approximately 55° CA before CTDC as the engine load decreases.

[0139] In contrast, as shown in Fig. As illustrated in Figure 12, the fuel amount of the third injection progressively decreases within the low-load segment of the third position, where all-cylinder operation is performed, and is further reduced at the engine load at which the operating state switches to cylinder deactivation. The fuel amount of the third injection is increased as the engine load further decreases.

[0140] Fig. Figure 13 illustrates a relationship between the total fuel quantity of the fuel quantities of the first to third injections in a combustion cycle and the engine load. Fig. As illustrated in Figure 13, the total fuel quantity is increased at the engine load at which the operating state switches to cylinder deactivation operation. Thus, the throttle loss in reduced-cylinder operation, ie, two-cylinder operation, is reduced, and stable SPCCI combustion is achieved or maintained even within the low-load range where an adverse effect on SPCCI combustion occurs, while making the air-fuel ratio A / F leaner.

[0141] Fig. Figure 14 illustrates an example of a relationship between the BMEP (corresponding to the engine load) and an exhaust valve closing timing (EVC) in positions 2 and 3. Fig. Figure 15 illustrates an example of a relationship between the BMEP (corresponding to the engine load) and an intake valve opening (IVO) timing in positions 2 and 3. Fig. Figure 16 illustrates an example of a relationship between the BMEP (corresponding to the engine load) and an additional rate of external EGR in positions 2 and 3.

[0142] As this is Fig. As illustrated in Figure 14, the EVC after the TDC in position 3 is earlier by about half the time of that in position 2.

[0143] On the other hand, as in Fig. 15, the IVO is the same in position 2 and position 3, including operation with reduced cylinders.

[0144] In addition, as stated in Fig. As illustrated in Figure 16, in layer 2, the additional rate of the external EGR decreases progressively up to a limit with layer 3. On the other hand, the additional rate is approximately 0% over the entire layer 3, in other words, the external EGR is not added. (Control process of an operation with cylinder deactivation of the engine)

[0145] Next, an operation control of the engine 1 executed by the ECU 10 will be described with reference to the flowchart of Fig. 17. First, at S21 after starting, the ECU 10 reads, which is in Fig. 4, the signals of the sensors SW1 to SW17, SW20 to SW24, SW31, SW51, SW101, SW102 and SW161. Next, at S22, the ECU 10 determines the operating state of the engine 1. Here, it is determined whether the operating range in position 3, which is Fig. 5, specifically the low load region (1)-1 (SPCCI λ>1 region), and if it is in position 3, it is determined whether the intake air temperature and the cooling water temperature are both above a given temperature. If not in position 3, or even in the case of position 3, if either the intake air temperature or the cooling water temperature is below the given temperature, the process returns to S21.

[0146] When the operating range is in position 3 and both the inlet air temperature and the cooling water temperature are above the given temperature, the process goes to S23.

[0147] At S23, it is determined whether the operating range is the reduced-cylinder operation range (cylinder deactivation operation). Whether it is in the reduced-cylinder operation range is determined by determining whether the BMEP is approximately 200 kPa or less. For example, the accelerator opening sensor SW12 and / or the throttle opening sensor SW17, etc., can be used.

[0148] Here, if it is determined that it is in the reduced-cylinder operating region, the process proceeds to S24. At S24, as described above, the ECU 10 outputs a control signal to the SCV 56 so that its opening ratio becomes approximately 0 to approximately 15%. Thus, the swirl ratio is approximately 4 to approximately 6.

[0149] Next, the reduced-cylinder operation is performed at S25. That is, the ECU 10 outputs a control signal to the fuel supply system 61 so that the fuel supply is stopped for a given two of the four cylinders. Furthermore, the ECU 10 injects fuel to the other two cylinders, which perform operation according to the total of three injection timings specified in Fig. 7 to 9, and the corresponding fuel quantities, which are shown in Fig. 10 to 12. Then the process returns to S21. (Other embodiments)

[0150] It should be noted that the teachings disclosed herein are not restricted or limited to application to the engine 1 having the above configuration. The configuration of the engine 1 can take various configurations.

[0151] Furthermore, the engine 1 may include a turbocharger instead of the mechanical booster or compressor 44. DESCRIPTION OF REFERENCE SYMBOLS 1 engine 10 ECU (control unit) 11 cylinders 17 Combustion chamber 171 crush area 25 Spark plug 3 pistons 31 cavity 401 primary passage (first inlet passage) 402 secondary passage (second inlet passage) 56 Vortex control valve (part that creates or generates a vortex) 6 Injection device

Claims

[1] Compression ignition engine (1), comprising: a plurality of cylinders (11), each formed with a combustion chamber (17); a spark plug (25) arranged in each of the combustion chambers (17); a fuel injection device (6) arranged to be directed into each combustion chamber (17); and a control unit (10) including a processor (101) connected to the spark plug (25) and the fuel injector (6) and configured to output a control signal to the spark plug (25) and the fuel injector (6), respectively, wherein when the engine (1) is operating at a load above a given load, the control unit (10) causes the engine (1) to operate with all cylinders by supplying fuel to all of the plurality of cylinders (11), and when the engine (1) is operating at a load below the given load, the control unit (10) causes the engine (1) to operate with reduced cylinders by supplying fuel to one or some of the plurality of cylinders (11), and in the reduced-cylinder operation, the control unit (10) controls the fuel injector (6) to inject fuel into one or some of the plurality of cylinders (11) to generate a mixture gas, controls the spark plug (25) to ignite the mixture gas, and causes the engine (1) to start, at an air-fuel ratio greater than a stoichiometric air-fuel ratio and at a large compression ratio above a given value, combustion by spark ignition (SI), in which the mixture gas is ignited by the spark plug (25) to burn by flame propagation, and then to perform combustion by compression ignition (CI), in which unburned mixture gas ignites by self-ignition, when the engine (1) is operating at a load below the given load, the control unit (10) switches an operating mode of the engine (1) between a first combustion mode in which the SI combustion and the CI combustion are carried out in this order at an air-fuel ratio substantially equal to a stoichiometric air-fuel ratio, and a second combustion mode in which the SI combustion and the CI combustion are carried out in this order at an air-fuel ratio greater than the stoichiometric air-fuel ratio, and when the engine (1) operates in the first combustion mode, the reduced cylinder operation is not performed. [2] An engine according to claim 1, wherein in reduced cylinder operation, the control unit (10) actuates an intake valve (21) and an exhaust valve (22) of each of a remainder of the plurality of cylinders (11). [3] An engine according to claim 1 or 2, wherein when a temperature of intake air into the combustion chamber (17) is below a given value and / or when a temperature of cooling water of the engine (1) is below a given value, the control unit (10) performs the all-cylinder operation instead of the reduced-cylinder operation. [4] An engine according to any preceding claim, further comprising a swirl control valve (56) configured to generate a swirl flow in each combustion chamber (17), wherein the control unit (10) controls the swirl control valve (56) to generate the swirl flow in the combustion chamber (17) in the reduced cylinder operation. [5] The motor according to claim 4, wherein a swirl ratio of the eddy current is about 2 or more. [6] Compression ignition engine (1), comprising: a plurality of cylinders (11), each formed with a combustion chamber (17); a spark plug (25) arranged in each combustion chamber (17); a fuel injection valve (6) arranged to be directed into each combustion chamber (17); an air-fuel ratio control valve (43) configured to control an air-fuel ratio of a mixture gas within the combustion chamber (17); and a control unit (10) connected to the spark plug (25), the fuel injection valve (6) and the air-fuel ratio control valve (43) and configured to output a control signal to the spark plug (25), the fuel injection valve (6) and the air-fuel ratio control valve (43), respectively, wherein the control unit (10) includes a processor (101) which is configured to execute or comprises: an engine load determination module (104) for determining whether an engine load is above a given load; an all-cylinder operation control module (105) for controlling the fuel injection valve (6) to inject fuel to all of the plurality of cylinders (11) when the engine load is determined to be above the given load; and a reduced cylinder operation control module (106) for, when the engine load is determined to be below the given load, controlling the fuel injection valve (6) to interrupt the injection of fuel to one or some of the plurality of cylinders (11), and wherein the reduced cylinder operation control module (106) comprises: an air-fuel ratio control submodule (107) configured to control the air-fuel ratio control device to make or set the air-fuel ratio greater than the stoichiometric air-fuel ratio; and a spark plug control submodule (108) configured to output a control signal to the spark plug (25) to perform ignition at a given ignition timing so that the mixed gas starts combustion by flame propagation and then unburned mixed gas self-ignites, the given ignition timing being stored in a memory (102), wherein: when the engine (1) is operating at a load below the given load, the control unit (10) switches an operating mode of the engine (1) between a first combustion mode in which the SI combustion and the CI combustion are carried out in this order at an air-fuel ratio substantially equal to a stoichiometric air-fuel ratio, and a second combustion mode in which the SI combustion and the CI combustion are carried out in this order at an air-fuel ratio greater than the stoichiometric air-fuel ratio, and when the engine (1) operates in the first combustion mode, the reduced cylinder operation is not performed. [7] A method for controlling a compression ignition engine having a plurality of cylinders (11), comprising the steps of: when the engine (1) is operating at a load above a given load, performing all-cylinder operation by supplying fuel to all of the plurality of cylinders (11), and when the engine (1) is operating at a load below the given load, performing reduced cylinder operation by supplying fuel to one or some of the plurality of cylinders (11), and in the reduced cylinder operation, injecting the fuel to one or some of the plurality of cylinders (11) to produce a mixture gas, and causing the engine (1) to start, at an air-fuel ratio greater than a stoichiometric air-fuel ratio and at a large compression ratio above a given value, spark-ignition combustion (SI) in which the mixture gas is ignited by a spark plug (25) to burn by flame propagation, and then perform compression-ignition combustion (CI) in which unburned mixture gas ignites by self-ignition, when the engine (1) is operating at a load below the given load, switching an operating mode of the engine (1) between a first combustion mode in which the SI combustion and the CI combustion are carried out in this order at an air-fuel ratio substantially equal to a stoichiometric air-fuel ratio, and a second combustion mode in which the SI combustion and the CI combustion are carried out in this order at an air-fuel ratio greater than the stoichiometric air-fuel ratio, wherein when the engine (1) operates in the first combustion mode, the reduced cylinder operation is not performed. [8] A control system for a compression ignition engine, comprising a control unit (10) configured to perform the steps of the method of claim 7. [9] A computer program product comprising computer-readable instructions which, when loaded onto and executed on a suitable system, can perform the steps of the method of claim 7.

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