JP2005054677A - Spark ignition type engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To obtain a fuel consumption improving effect by a lean combustion, a pumping loss reduction, etc. while assuring sufficiently an exhaust gas purifying performance with a simple constitution. <P>SOLUTION: A spark ignition type engine includes an operation mode control means for performing a control of a special operation mode for burning together with newly supplied fuel in following cylinders 2B, 2C by burning with an air fuel ratio of preceding cylinders 2A, 2D as a lean air fuel ratio larger than a theoretical air fuel ratio, and introducing a burned gas of the lean air fuel ratio from the preceding cylinder 2A, 2D to the following cylinders 2B, 2C. A gas discharging valve 32b is provided at an upstream end of a gas passage between the cylinders connected to the preceding cylinders 2A, 2D, a gas introducing valve 31b is provided at a downstream end of the gas passage between the cylinders connected to the following cylinders 2B, 2C, and a valve opening timing of the gas introducing valve 31b at the special operation mode control time is set to a time later than the valve opening timing of the gas discharging valve 32b. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

  The present invention relates to a multi-cylinder spark ignition engine in which the combustion cycle of each cylinder is set to have a predetermined phase difference.

  2. Description of the Related Art Conventionally, in a spark ignition engine, a technique for improving fuel consumption by performing combustion in a state where the air-fuel ratio of the air-fuel mixture in each cylinder is set to a lean air-fuel ratio larger than the stoichiometric air-fuel ratio is known. It has a fuel injection valve that directly injects fuel into the room, and in low-speed and low-load areas, etc., super lean combustion is realized by injecting fuel from the fuel injection valve in the compression stroke and performing stratified combustion. Those are known (for example, see Patent Document 1).

  In such an engine, an ordinary three-way catalyst (a catalyst having a high purification performance in the vicinity of the theoretical air-fuel ratio with respect to HC, CO, and NOx) is sufficient for NOx during lean operation as a catalyst for exhaust gas purification. Therefore, as shown in Patent Document 1, there is provided a lean NOx catalyst having a predetermined capacity for adsorbing NOx in an oxygen-excess atmosphere and detaching and reducing NOx in an oxygen concentration-reduced atmosphere. . When the lean NOx catalyst is used, if the NOx adsorption amount of the lean NOx catalyst increases during the lean operation, for example, as shown in Patent Document 1, additional fuel is injected during the expansion stroke in addition to the main combustion. The air-fuel ratio of the exhaust gas is enriched and CO is generated, thereby promoting NOx removal and reduction.

As another method for improving fuel efficiency, for example, as shown in Patent Document 2 below, by leaving a large amount of burned gas in the combustion chamber, the combustion chamber can be moved in the same manner as a diesel engine at the end of the compression stroke. The air-fuel mixture is self-ignited at high temperature and high pressure (compression self-ignition). When such compression self-ignition is performed, combustion occurs all at once in the combustion chamber, so it does not contribute to work. It is possible to avoid combustion, which is advantageous for improving fuel consumption, and it is possible to prevent the temperature in the combustion chamber from becoming locally high and to suppress generation of NOx.
Japanese Patent Laid-Open No. 10-29836 JP-A-10-266878

  In an engine that performs the conventional lean operation as shown in Patent Document 1, it is necessary to provide the lean NOx catalyst in the exhaust passage in order to ensure NOx purification performance during the lean operation, which is disadvantageous in terms of cost. . In order to maintain the purification performance of the lean NOx catalyst, as described above, NOx is released and reduced when the NOx adsorption amount increases, so that the air-fuel ratio is temporarily enriched by supplying additional fuel or the like. There is a need. Furthermore, when the fuel used contains a large amount of sulfur, in order to eliminate the sulfur poisoning of the lean NOx catalyst, a regeneration process consisting of heating the catalyst and supplying a reducing material is required. The effect is inevitable. Moreover, if the air-fuel ratio of the air-fuel mixture becomes leaner than a certain level, the combustion speed becomes too slow and combustion close to the end does not contribute to work, so there is a limit to fuel efficiency improvement by leaning in stratified combustion .

  On the other hand, as shown in Patent Document 2, in a normal spark ignition type gasoline engine, when it is configured to perform compression self-ignition in order to obtain an improvement effect of fuel consumption and an NOx suppression effect, the compression is performed. There is a problem in that special measures are required to greatly increase the temperature or pressure in the combustion chamber near the top dead center so as to obtain an environment of compression self-ignition. In addition, there has been a problem that it is difficult to reliably perform compression self-ignition even if the above-mentioned device is elaborated in an ordinary spark ignition type gasoline engine.

  For this reason, the applicant of the present application is a multi-cylinder engine that performs a cycle consisting of intake, compression, expansion, and exhaust strokes, and a pair of cylinders in which the exhaust stroke and the intake stroke overlap at least in a partial load region on the low load and low rotation side. The burned gas discharged from the preceding cylinder in the exhaust stroke is introduced as it is into the subsequent cylinder in the intake stroke via the inter-cylinder gas passage, and the three-way catalyst provides the gas discharged from this succeeding cylinder By setting the two-cylinder connection state leading to the exhaust passage, combustion is performed in a state where the preceding cylinder has a lean air-fuel ratio that is significantly larger than the stoichiometric air-fuel ratio, and in the subsequent cylinder, the lean air-fuel ratio introduced from the preceding cylinder is performed. Has been developed (Japanese Patent Application No. 2002-024548).

  According to the above configuration, at least in the low-load and low-speed region of the engine, the preceding cylinder performs combustion at a lean air-fuel ratio, thereby improving the thermal efficiency and reducing the pumping loss, thereby obtaining a remarkable fuel efficiency improvement effect. In the succeeding cylinder, fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinder, and combustion is performed in a state where the stoichiometric air-fuel ratio is set, so that a fuel efficiency effect by reducing pumping loss can be obtained. Moreover, since only the stoichiometric burned gas discharged from the subsequent cylinders is guided to the exhaust passage provided with the three-way catalyst, sufficient exhaust purification performance is ensured with only the three-way catalyst, and no lean NOx catalyst is required. Become. Further, since the high-temperature burned gas is introduced into the succeeding cylinder from the preceding cylinder by setting the two-cylinder connection state as described above, the combustion chamber of the succeeding cylinder is self-ignited at a high temperature and high pressure in the latter half of the compression stroke. As a result, the air-fuel mixture can be combusted all over the combustion chamber, so that slow combustion that does not contribute to work can be avoided, which is further advantageous in improving fuel consumption.

  The present invention has been made in consideration of the above-described problems, and is a spark ignition engine capable of obtaining fuel efficiency improvement effects such as lean combustion and reduction of pumping loss while ensuring sufficient exhaust purification performance with a simple configuration. Is to provide.

  The invention according to claim 1 is a multi-cylinder spark ignition engine in which the combustion cycle of each cylinder is set to have a predetermined phase difference, and the exhaust stroke and the intake stroke are reduced in a partial load region of the engine. The burned gas discharged from the preceding cylinder in the exhaust stroke between a pair of matching cylinders is directly introduced into the succeeding cylinder in the intake stroke through the inter-cylinder gas passage, and the burned gas discharged from the succeeding cylinder is Combustion is performed with the air-fuel ratio of the preceding cylinder set to a lean air-fuel ratio larger than the stoichiometric air-fuel ratio while the two-cylinder connection state is led to the exhaust passage, and the burned gas having the lean air-fuel ratio is transferred from the preceding cylinder to the succeeding cylinder. An upstream end of an inter-cylinder gas passage connected to the preceding cylinder is provided with an operation mode control means for performing control of a special operation mode for burning in the subsequent cylinder together with the newly supplied fuel. And a gas introduction valve at the downstream end of the inter-cylinder gas passage connected to the succeeding cylinder, and the timing of opening the gas introduction valve during the control of the special operation mode is It is set at a time later than the valve opening time.

  According to a second aspect of the present invention, in the spark ignition type engine according to the first aspect, the opening timing of the gas introduction valve at the time of control in the special operation mode is set at or near the exhaust top dead center of the subsequent cylinder. The opening timing of the gas outlet valve is set earlier than the exhaust top dead center of the subsequent cylinder.

  According to a third aspect of the present invention, in the spark ignition engine according to the first or second aspect, the exhaust valve for exhausting the exhaust gas of the subsequent cylinder to the exhaust passage and the gas introduction valve are adjacent to the subsequent cylinder. Arranged.

  According to a fourth aspect of the present invention, in the spark ignition engine according to any one of the first to third aspects of the present invention, the closing timing of the gas introduction valve is determined when the gas introduction valve is closed during the control in the special operation mode. The time is set later than the time.

  According to a fifth aspect of the present invention, in the spark ignition type engine according to the fourth aspect, the closing timing of the gas outlet valve at the time of control in the special operation mode is set at or near the exhaust top dead center of the preceding cylinder. The closing timing of the gas introduction valve is set to a timing later than the exhaust top dead center of the preceding cylinder.

  The invention according to claim 6 is the spark ignition engine according to any one of claims 1 to 5, wherein the overlap period between the exhaust valve and the gas introduction valve of the subsequent cylinder at the time of control in the special operation mode. Is set to 5 ° or less in crank angle.

  The invention according to claim 7 is the spark ignition engine according to any one of claims 1 to 6, wherein the gas outlet valve and the gas provided at the upstream end and the downstream end of the inter-cylinder gas passage Each inlet valve is composed of poppet valves of the same diameter, and the set load of the valve spring that biases the gas inlet valve in the closing direction is higher than the set load of the valve spring that biases the gas outlet valve in the closing direction. Is set.

  In the invention according to claim 1, the gas provided at the downstream end of the inter-cylinder gas passage in a two-cylinder connection state in which the burned gas discharged from the preceding cylinder is introduced into the subsequent cylinder through the inter-cylinder gas passage. By setting the opening timing of the introduction valve to a relatively late timing and shortening the overlap period between the exhaust valve and the gas introduction valve that are in the closed state near the exhaust top dead center of the subsequent cylinder, Since it is possible to effectively prevent the burned gas introduced into the succeeding cylinder from the gas passage from being blown out from the exhaust port to the exhaust passage, it is possible to ensure combustibility in the succeeding cylinder. In addition, by setting the opening timing of the gas outlet valve provided on the preceding cylinder side to a relatively early timing, the flow resistance of the burned gas derived from the preceding cylinder into the inter-cylinder gas passage is increased. There is an advantage that the burned gas can flow smoothly.

  In the invention according to claim 2, the gas provided at the downstream end of the inter-cylinder gas passage in a two-cylinder connection state in which the burned gas discharged from the preceding cylinder is introduced into the subsequent cylinder through the inter-cylinder gas passage. Since the opening timing of the introduction valve is set at or near the exhaust top dead center of the subsequent cylinder, the burned gas introduced into the subsequent cylinder from the inter-cylinder gas passage is exhausted through the exhaust port during the exhaust stroke. Blowing inward can be prevented more effectively, and the gas introduction valve (poppet valve) provided above the piston head is not lifted near the top dead center of the piston. The piston top dead center position can be brought close to the gas introduction valve, and the geometric compression ratio of the succeeding cylinder is increased correspondingly, so that the succeeding cylinder can be surely compressed and self-ignited. In addition, since the opening timing of the gas outlet valve provided at the upstream end of the inter-cylinder gas passage is set earlier than the exhaust top dead center of the succeeding cylinder than the opening timing of the gas introducing valve, It is possible to effectively prevent the flow resistance of the burned gas led out into the inter-cylinder gas passage from increasing effectively, and the burned gas can flow smoothly.

  When the exhaust valve for discharging the exhaust gas of the succeeding cylinder to the exhaust passage and the gas introduction valve are arranged adjacent to each other in the succeeding cylinder as in the invention according to claim 3, the inter-cylinder gas passage From the above, the burned gas introduced into the succeeding cylinder easily blows through the exhaust port to the exhaust passage side, so that the burned gas discharged from the preceding cylinder as described above passes to the succeeding cylinder via the inter-cylinder gas passage. In the state where the two cylinders are connected, the opening timing of the gas introduction valve provided at the downstream end of the inter-cylinder gas passage is determined from the opening timing of the gas outlet valve provided at the upstream end of the inter-cylinder gas passage. The burned gas introduced into the succeeding cylinder is reduced by shortening the overlap period between the exhaust valve that is stopped near the exhaust top dead center of the succeeding cylinder and the gas introducing valve. Significant effect of preventing blow-off to the exhaust passage side There is an advantage that is obtained.

  In the invention according to claim 4, the gas provided at the downstream end of the inter-cylinder gas passage in a two-cylinder connection state in which the burned gas discharged from the preceding cylinder is introduced into the subsequent cylinder through the inter-cylinder gas passage. By making the introduction valve relatively closed, the flow loss of the burned gas introduced into the succeeding cylinder through the inter-cylinder gas passage is suppressed, and the burned gas is smoothly introduced into the succeeding cylinder. Thus, the combustibility of the succeeding cylinder can be improved, and the gas outlet valve on the preceding cylinder side is closed relatively quickly, thereby reducing the internal EGR amount of the preceding cylinder and reducing the amount of fresh air with respect to the preceding cylinder. An introduction amount can be secured.

  In the invention according to claim 5, the gas provided at the upstream end of the inter-cylinder gas passage in a two-cylinder connection state in which the burned gas discharged from the preceding cylinder is introduced into the subsequent cylinder through the inter-cylinder gas passage. By setting the closing timing of the lead-out valve at or near the exhaust top dead center of the preceding cylinder, the internal EGR amount of the preceding cylinder can be effectively reduced and a sufficient amount of fresh air introduced into the preceding cylinder can be secured. And the subsequent cylinder via the inter-cylinder gas passage by setting the closing timing of the gas introduction valve provided at the downstream end of the inter-cylinder gas passage to a time later than the exhaust top dead center of the preceding cylinder. The flow loss of the burned gas introduced into the inside can be effectively suppressed, and the burnt gas can be sufficiently introduced into the succeeding cylinder to reliably improve the combustibility of the succeeding cylinder.

  In the invention according to claim 6, the exhaust valve and the gas introduction valve of the subsequent cylinder overlap in a two-cylinder connection state in which the burned gas discharged from the preceding cylinder is introduced to the subsequent cylinder through the inter-cylinder gas passage. By setting the period of the state to be substantially 0, it is possible to suppress the burned gas introduced into the succeeding cylinder from blowing into the exhaust passage as much as possible.

  The invention according to claim 7 is the burned gas introduced into the inter-cylinder gas passage in a two-cylinder connection state in which the burned gas discharged from the preceding cylinder is introduced into the subsequent cylinder through the inter-cylinder gas passage. This effectively prevents the gas introduction valve from opening due to the pressure of the gas acting on the gas introduction valve, and the burned gas is introduced into the succeeding cylinder earlier than the set time. There is an advantage that it is possible to suppress the blow-by of the burned gas as much as possible.

  FIG. 1 shows a schematic configuration of an engine to which the present invention is applied, and FIG. 2 schematically shows a structure of one cylinder of an engine body 1 and intake / exhaust valves provided for the cylinder. In these drawings, the engine body 1 has a plurality of cylinders, and in the illustrated embodiment, has four cylinders 2A to 2D. A piston 3 is fitted into each of the cylinders 2 </ b> A to 2 </ b> D, and a combustion chamber 4 is formed above the piston 3. A spark plug 7 is provided at the top of the combustion chamber 4 of each cylinder 2 </ b> A to 2 </ b> D, and the tip of the plug faces the combustion chamber 4. An ignition circuit 8 capable of controlling the ignition timing by electronic control is connected to the spark plug 7.

  A fuel injection valve 9 that directly injects fuel into the combustion chamber 4 is provided at a side portion of the combustion chamber 4. The fuel injection valve 9 includes a needle valve and a solenoid (not shown). When a pulse signal is input, the fuel injection valve 9 is driven for a time corresponding to the pulse width at the pulse input timing to open the valve. It is comprised so that the quantity of fuel according to may be injected. The fuel injection valve 9 is supplied with fuel by a fuel pump (not shown) via a fuel supply passage and the like, and a fuel supply system is provided so that a fuel pressure higher than the pressure in the combustion chamber in the compression stroke can be applied. It is configured.

  Each cylinder 2A to 2D is configured to perform a cycle including intake, compression, expansion, and exhaust strokes with a predetermined phase difference. In the case of a four-cylinder engine, the first cylinder from one end in the cylinder row direction 2A, 2nd cylinder 2B, 3rd cylinder 2C, 4th cylinder 2D, as shown in FIG. 5, the cycle is in the order of 1st cylinder 2A, 3rd cylinder 2C, 4th cylinder 2D, 2nd cylinder 2B. The crank angle is performed with a phase difference of 180 °. In FIG. 5, EX is an exhaust stroke, IN is an intake stroke, F is fuel injection, S is forced ignition, and a star mark in the drawing indicates that compression ignition is performed.

  Between a pair of cylinders in which the exhaust stroke and the intake stroke coincide with each other, from the cylinder on the exhaust stroke side when the exhaust stroke and the intake stroke coincide with each other and exactly overlap each other (this is referred to as a preceding cylinder in this specification) An inter-cylinder gas passage 22 is provided so that the burned gas can be guided as it is to a cylinder on the intake stroke side (referred to as a subsequent cylinder in the present specification). In the four-cylinder engine of this embodiment, as shown in FIG. 5, the exhaust stroke (EX) of the first cylinder 2A and the intake stroke (IN) of the second cylinder 2B overlap, and the exhaust stroke (EX) of the fourth cylinder 2D. ) And the intake stroke (IN) of the third cylinder 2C overlap, so that the first cylinder 2A and the second cylinder 2B, and the fourth cylinder 2D and the third cylinder 2C form a pair, respectively, and the first cylinder 2A and the fourth cylinder The cylinder 2D is the preceding cylinder, the second cylinder 2B, and the third cylinder 2C are the subsequent cylinders.

  Specifically, the intake / exhaust port of each cylinder and the intake passage, exhaust passage, and inter-cylinder gas passage connected thereto are configured as follows. The first cylinder 2A and the fourth cylinder 2D, which are the preceding cylinders, respectively include an intake port 11 for introducing fresh air, and a first exhaust port 12a for sending burned gas (exhaust gas) to the exhaust passage. A second exhaust port 12b for leading the burned gas to the subsequent cylinder is provided. The second cylinder 2B and the third cylinder 2C, which are the subsequent cylinders, respectively, have a first intake port 11a for introducing fresh air and a second intake port for introducing burned gas from the preceding cylinder. 11b and an exhaust port 12 for sending burned gas to the exhaust passage.

  In the example shown in FIG. 1, two intake ports 11 in the preceding cylinders 2A and 2D and two first intake ports 11a in the succeeding cylinders 2B and 2C are provided in parallel in one half of the combustion chamber. ing. Further, the first exhaust port 12a and the second exhaust port 12b in the preceding cylinders 2A and 2D and the second intake port 11b and the exhaust port 12 in the subsequent cylinders 2B and 2C are provided in parallel in the other half of the combustion chamber. ing. That is, an exhaust valve 32 that opens and closes the exhaust port 12 and a gas introduction valve 31b that opens and closes the second intake port 11b are disposed adjacent to each other in the other half of the succeeding cylinders 2B and 2C, as will be described later. Has been.

  A downstream end of a branch intake passage 16 for each cylinder in the intake passage 15 is connected to the intake port 11 in the preceding cylinders 2A and 2D and the first intake port 11a in the subsequent cylinders 2B and 2C. In the vicinity of the downstream end of each branch intake passage 16, a multiple throttle valve 17 that is linked to each other via a common shaft is provided. This multiple throttle valve 17 is driven by an actuator 18 in accordance with a control signal, The intake air amount is adjusted. Note that an air flow sensor 19 that detects an intake air flow rate is provided in a common intake passage upstream of the collecting portion in the intake passage 15.

  An upstream end portion of a branch exhaust passage 21 for each cylinder in the exhaust passage 20 is connected to the first exhaust port 12a in the preceding cylinders 2A and 2D and the exhaust port 12 in the subsequent cylinders 2B and 2C. Further, an inter-cylinder gas passage 22 is provided between the first cylinder 2A and the second cylinder 2B and between the third cylinder 2C and the fourth cylinder 2D, respectively, and the first and fourth cylinders which are the preceding cylinders. The upstream end portion of the inter-cylinder gas passage 22 is connected to the second exhaust ports 12b of 2A and 2D, and the inter-cylinder gas passage 22 is connected to the second intake ports 11b of the second and third cylinders 2B and 2C as the subsequent cylinders. Are connected at the downstream end.

The inter-cylinder gas passage 22 is provided with a linear O ₂ sensor 25 whose output linearly changes in accordance with the oxygen concentration, and for the preceding cylinders 2A and 2D that have a predetermined lean air-fuel ratio based on the output. The fuel injection amount is feedback controlled.

An O ₂ sensor 23 that detects the air-fuel ratio by detecting the oxygen concentration in the exhaust gas is provided at the downstream of the branch exhaust passage 21 in the exhaust passage 20. The O ₂ sensor 23 is a λO ₂ sensor whose output changes suddenly in the vicinity of the theoretical air-fuel ratio. Based on the output of the λO ₂ sensor 23, the succeeding cylinders 2B and 2C (the preceding cylinders 2A and 2D when each cylinder is in an independent state) The fuel injection amount with respect to (including) is feedback-controlled. Further, a three-way catalyst 24 for exhaust purification is provided in the exhaust passage 20 downstream of the O ₂ sensor 23. As is generally known, the three-way catalyst 24 has high purification performance for HC, CO, and NOx when the air-fuel ratio of the exhaust gas is near the stoichiometric air-fuel ratio (that is, the excess air ratio λ is 1). It is the catalyst shown.

  The valves for opening and closing the intake / exhaust ports of each cylinder and the valve operating mechanism for these valves are as follows. The intake port 11, the first exhaust port 12a and the second exhaust port 12b in the preceding cylinders 2A and 2D are provided with an intake valve 31, an exhaust valve 32a and a gas outlet valve 32b, respectively, and the first intake in the subsequent cylinders 2B and 2C. The port 11a, the second intake port 11b, and the exhaust port 12 are provided with an intake valve 31a, a gas introduction valve 31b, and an exhaust valve 32, respectively.

  Each of the valves 31, 32a, 32b, 31a, 31b, and 32 is a poppet valve that is urged in the closing direction by a valve spring (not shown). When each cylinder is in the intake stroke or the exhaust stroke, the camshaft 33 , 34 is pushed down by the drive cams provided in the open state, but the opening / closing timing is not necessarily limited to top dead center or bottom dead center, and a predetermined crank angle from top dead center or bottom dead center as required. It is set at a time shifted by only. The set load of the valve spring that urges the gas introduction valve 31b provided at the downstream end of the inter-cylinder gas passage 22 in the closing direction, that is, the urging force of the valve spring when the gas introduction valve 31b is closed is It is set to a value higher than the set load of the valve spring that biases the gas outlet valve 32b provided at the upstream end of the inter-gas passage 22 in the closing direction. Note that the diameters of the poppet valves constituting the gas introduction valve 31b and the gas outlet valve 32b are set to the same diameter, and the set load of the valve springs may be the same considering the specifications of both poppet valves. In the embodiment of the present invention, the set load of the valve spring on the gas introduction valve 31b side is set relatively higher than that on the gas outlet valve 32b side.

  Further, among the above valves, for the second exhaust valve 32a, the gas outlet valve 32b, the first intake valve 31 and the gas introduction valve 31b, a valve stop mechanism 35 for switching these valves between an operating state and a stopped state. Is provided. The valve stop mechanism 35 has been known in the art and will not be shown in detail. For example, a hydraulic chamber that can supply and discharge hydraulic oil is provided in a tappet interposed between a cam and a valve shaft. When hydraulic oil is supplied to the hydraulic chamber, the cam operation is transmitted to the valve and the valve is opened and closed. When hydraulic oil is discharged from the hydraulic chamber, transmission of power from the cam to the valve is interrupted. The valve is configured to be stopped.

  A first control valve 37 is provided in the hydraulic oil supply / discharge passage 36 to the valve stop mechanism 35 of the intake valve 31a on the subsequent cylinders 2B, 2C side and the exhaust valve 32a on the preceding cylinders 2A, 2D side. A second control valve 39 is provided in the hydraulic oil supply / discharge passage 38 with respect to the valve stop mechanism 35 of the inlet valve 31b and the gas outlet valve 32b (see FIG. 3).

FIG. 3 shows the configuration of the drive and control system in this embodiment. In this figure, an engine control ECU (control unit) 40 composed of a microcomputer or the like receives signals from an air flow sensor 19, an O ₂ sensor 23 and a linear O ₂ sensor 25, and determines an operating state. Therefore, signals from an engine speed sensor 47 for detecting the engine speed and an accelerator position sensor 48 for detecting an accelerator position (accelerator pedal depression amount) are inputted. Control signals are output from the ECU 40 to the ignition circuit 8, the fuel injection valves 9, the actuator 18 of the multiple throttle valve 17, and the first and second control valves 37 and 39.

  The ECU 40 constitutes a control means for performing combustion while keeping the gas flow path in a two-cylinder connection state (see FIG. 6) at least in a partial load region on the low load and low rotation side of the engine. Means 41, valve stop mechanism control means 42, intake air amount control means 43, combustion control means 44 including fuel injection control means 45 and ignition control means 46 are provided.

  The operating state discriminating means 41 is based on the detection signal of the engine operating state (engine speed and engine load) output from the rotational speed sensor 47, the accelerator opening sensor 48, etc., and the operating state is low as shown in FIG. It is determined whether the load is in the partial load region A on the low rotation side, the full load region B on the high load side or on the high rotation side, or the idle operation region C. When the engine is in the partial load region A in the warm-up state), the combustion control is performed in the special operation mode in which the two-cylinder connection state is established, and when the engine is in the full load region B or the idle operation region C, Combustion control is performed in a normal operation mode in which the cylinder is independent.

  The valve stop mechanism control means 42 controls each of the valve stop mechanisms 35 as follows by controlling the control valves 37 and 39 in accordance with the determination result of the special operation mode and the normal operation mode.

Special operation mode: The leading cylinder exhaust valve 32a and the trailing cylinder intake valve 31a are stopped.
Gas derivation valve 32b and gas introduction valve 31b are in operating state Normal operation mode: Leading cylinder exhaust valve 32a and trailing cylinder intake valve 31a are in operating state
The gas outlet valve 32b and the gas introduction valve 31b are stopped. The intake air amount control means 43 controls the opening degree (throttle opening degree) of the throttle valve 17 by controlling the actuator 18, and according to the operating state. A target intake air amount is obtained from a map or the like, and the throttle opening is controlled according to the target intake air amount. In this case, in the special operation mode, as will be described later, in the succeeding cylinders 2B and 2C, excess air in the gas introduced from the preceding cylinder is newly supplied in a state where the intake air introduction from the branch intake passage 16 is blocked. Since the combustion is performed while the ratio to the fuel is the stoichiometric air-fuel ratio (hereinafter referred to as a substantial stoichiometric air-fuel ratio), the amount of fuel necessary for the combustion of the fuel corresponding to the required torque for the preceding and succeeding two cylinders The throttle opening is adjusted so that air (the amount of air that is the stoichiometric air-fuel ratio with respect to the amount of fuel for two cylinders) is supplied to the preceding cylinders 2A and 2D.

  The combustion control unit 44 includes a fuel injection control unit 45 and an ignition control unit 46. The fuel injection control unit 45 causes the fuel injection amount and the injection timing from the fuel injection valves 9 provided in the respective cylinders 2A to 2D. Is controlled according to the operating state of the engine, and the ignition control means 46 controls the ignition timing and the ignition stop according to the operating state. In particular, the combustion control (fuel injection control and ignition control) state is changed depending on whether the operation state is the special operation mode or the normal operation mode.

  That is, in the partial load region A in which the engine is controlled in the special operation mode by the operation mode control means including the valve stop mechanism control means 42 and the like, the air-fuel ratio is set to the theoretical empty for the preceding cylinders 2A and 2D. The fuel injection amount is controlled to be a lean air-fuel ratio larger than the fuel ratio, preferably about twice or more than the theoretical air-fuel ratio, and fuel is injected in the compression stroke so that the mixture is stratified. And the ignition timing is set so that forced ignition is performed near the compression top dead center. On the other hand, for the succeeding cylinders 2B and 2C, fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinders 2A and 2D, and the fuel injection amount is set so as to obtain a substantial stoichiometric air-fuel ratio. In addition to being controlled, the injection timing is set so as to inject fuel in the intake stroke, and combustion by compression self-ignition or forced ignition is performed according to the operating state.

  Further, when the engine is in the full load region B or the idle operation region C, the fuel injection amount is controlled so that the air-fuel ratio of each of the cylinders 2A to 2D is equal to or lower than the theoretical air-fuel ratio, and combustion in the normal operation mode is performed. The control is executed, for example, the air-fuel ratio of each cylinder 2A to 2D is set to the stoichiometric air-fuel ratio in the most region in the normal operation mode, and the air-fuel ratio of each cylinder 2A to 2D in the operating region near the maximum load. More rich control is executed. In this case, the injection timing is set so that fuel is injected into each cylinder 2A to 2D in the intake stroke and the air-fuel mixture is made uniform, and forced ignition is performed in each cylinder 2A to 2D. Be controlled.

  The operation of the apparatus of the present embodiment as described above will be described with reference to FIGS. In the special operation mode, as described above, the exhaust valves 32a of the preceding cylinders 2A and 2D and the intake valves 31a of the succeeding cylinders 2B and 2C are stopped, and the gas outlet valve 32b provided at the upstream end of the inter-cylinder gas passage 22 and When the gas introduction valve 31b provided at the downstream end of the inter-cylinder gas passage 22 is activated, the substantial fresh air and gas flow paths are brought into a two-cylinder connection state as shown in FIG. The burned gas discharged from the preceding cylinders 2A and 2D is directly introduced into the succeeding cylinders 2B and 2C through the inter-cylinder gas passage 22, and only the exhaust gas discharged from the succeeding cylinders 2B and 2C is discharged into the exhaust passage. Will be led to 20.

In the two-cylinder connected state, fresh air is introduced into the preceding cylinders 2A and 2D from the intake passage 15 during the intake stroke (arrow a in FIG. 6), and is detected by the linear O ₂ sensor 25 in the preceding cylinders 2A and 2D. The fuel injection amount is feedback controlled so that the air-fuel ratio becomes a super-lean air-fuel ratio approximately twice or more than the stoichiometric air-fuel ratio, fuel is injected in the compression stroke, and ignition is performed at a predetermined ignition timing. Thus, stratified combustion is performed at an ultra lean air-fuel ratio (see FIG. 5).

  Thereafter, during a period in which the intake strokes of the preceding cylinders 2A and 2D coincide with the exhaust strokes of the succeeding cylinders 2B and 2C, the burned gas discharged from the preceding cylinders 2A and 2D passes through the gas passage 22 to the succeeding cylinders 2B and 2C. It is introduced (the white arrow in FIG. 5 and the arrow b in FIG. 6). In the succeeding cylinders 2B and 2C, fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinders 2A and 2D, and the fuel injection amount is controlled so as to obtain a substantial stoichiometric air-fuel ratio. After the fuel is injected in the intake stroke, the pressure and temperature in the combustion chamber rise near the top dead center of the compression stroke, and compression self-ignition is performed.

  In this case, as shown in FIG. 7, the opening timing α1 of the gas introduction valve 31b provided at the downstream end of the inter-cylinder gas passage 22 is at or near the exhaust top dead center TDC of the succeeding cylinders 2B and 2C. The opening timing β1 of the gas outlet valve 32b provided at the upstream end of the inter-cylinder gas passage 22 is set to a predetermined crank angle, for example, 50 from the exhaust top dead center TDC of the succeeding cylinders 2B and 2C. By being set only before CA (crank angle), the valve opening timing α1 of the gas introduction valve 31b during the control of the special operation mode is set to a timing later than the valve opening timing β1 of the gas outlet valve 32b. Yes.

In the embodiment of the present invention, in order to improve the accuracy of air-fuel ratio control of the preceding cylinders 2A and 2D, a linear O ₂ sensor 25 is provided in the inter-cylinder gas passage 22 to feed back the fuel injection amounts of the preceding cylinders 2A and 2D. Although controlled, the linear O ₂ sensor 25 may be omitted. That is, the fuel injection amount of the preceding cylinders 2A and 2D is determined from the air flow sensor 19, the O ₂ sensor 23, the rotation speed sensor 47, the accelerator opening sensor 48, and the ECU 40, and the air-fuel ratio set in advance according to the engine operating state. Thus, the fuel injection amounts of the preceding cylinders 2A and 2D corresponding to the intake air amount are determined (open control), and the succeeding cylinders 2B and 2C have the stoichiometric air-fuel ratio based on the output of the O ₂ sensor 23. Thus, the fuel injection amount may be feedback controlled. Further, the fuel injection amounts of both the preceding cylinders 2A and 2D and the succeeding cylinders 2B and 2C may be determined based on the output of the O ₂ sensor 23.

  On the other hand, the closing timing γ2 of the exhaust valve 32 provided in the succeeding cylinders 2B, 2C is the exhaust top dead center TDC of the succeeding cylinders 2B, 2C or the vicinity thereof, for example, a timing of about 4 ° CA after the exhaust top dead center TDC. And the overlap period OB between the exhaust valve 32 and the gas introduction valve 31b provided adjacent to the combustion chambers of the succeeding cylinders 2B and 2C is set within 5 ° CA in crank angle, A period in which the exhaust valve 32 and the gas introduction valve 31b are substantially overlapped is substantially zero.

  The valve closing timing β2 of the gas outlet valve 32b provided at the upstream end of the inter-cylinder gas passage 22 is set at or near the exhaust top dead center TDC of the preceding cylinders 2A and 2D, and the inter-cylinder gas. The valve closing timing α2 of the gas introduction valve 31b provided at the downstream end of the passage 22 is set after a predetermined crank angle, for example, 54 ° CA, after the exhaust top dead center TDC of the preceding cylinders 2A and 2D. Thus, the closing timing α2 of the gas introduction valve 31b at the time of control in the special operation mode is set to a timing later than the closing timing β2 of the gas outlet valve 32b.

  As shown in FIG. 8, the lift characteristics of the gas inlet valve 31b and the gas outlet valve 32b are as follows. The opening side buffer portion L1 and the closing side buffer portion whose valve opening changes at a constant speed when the valve is opened and when the valve is closed. It has L2, and it is set so that it may have the head part H between these buffer parts L1, L2. In this embodiment, when the valve is opened, a boundary point between the opening side buffer portion L1 and the lift portion H, for example, the lift amount of the gas inlet valve 31b and the gas outlet valve 32b is about 0.4 mm, and the lift speed changes suddenly. Specifically, the time when the lift speed is switched from the constant speed section to the positive acceleration section is defined as the valve opening timings α1 and β1, and the lift speed is changed from the positive acceleration section (closed-side buffer section L2) to the constant speed section (lift section H). ) Is defined as the valve closing timings α2 and β2.

  Since the high-temperature burned gas discharged from the preceding cylinders 2A and 2D is introduced into the succeeding cylinders 2B and 2C through the inter-cylinder gas passage 22 as described above, the intake stroke is performed in the succeeding cylinders 2B and 2C. In this state, the temperature in the combustion chamber can be effectively increased, and from this state, by further increasing the pressure and temperature in the compression stroke, the air-fuel mixture is sufficiently compressed and self-ignited near the top dead center of the compression stroke. The temperature in the combustion chamber can be increased to the extent that it is obtained. In addition, the burned gas derived from the preceding cylinders 2A and 2D is sufficiently mixed and evenly distributed until it is introduced into the succeeding cylinders 2B and 2C, and injected into the succeeding cylinders 2B and 2C in the intake stroke. Since the burned fuel is also dispersed throughout the combustion chamber until the end of the compression stroke, an air-fuel mixture distribution state that satisfies the ideal simultaneous compression ignition condition can be obtained.

  Therefore, the subsequent cylinders 2B and 2C contain a large amount of burned gas components equivalent to EGR gas and the combustion is rapidly performed by simultaneous compression ignition even under the condition that the air-fuel ratio is lean. The thermal efficiency of the engine will be greatly improved. That is, in the preceding cylinders 2A and 2D, the heat efficiency is increased and the pumping loss is reduced by the stratified combustion in the ultra-lean state, and in the succeeding cylinders 2B and 2C, the pumping loss reducing effect is obtained as in the preceding cylinders 2A and 2D. Since the compression efficiency is performed in a uniform air-fuel mixture distribution state, the thermal efficiency is increased, and the fuel efficiency is greatly improved by these actions. Further, since the compression self-ignition in the succeeding cylinders 2B and 2C is achieved by using the temperature of the burned gas derived from the preceding cylinders 2A and 2D, a special heating means is used or the compression ratio of the engine is extremely reduced. There is an advantage that the compression self-ignition can be performed over a wide operating range without adopting a configuration such as increasing the speed.

  Further, the preceding cylinders 2A and 2D have a lean air-fuel ratio that is approximately twice or more than the theoretical air-fuel ratio, so that the amount of NOx generated is suppressed to be relatively small. In the succeeding cylinders 2B and 2C, the preceding cylinders 2A and 2D Since the burned gas is introduced from this state, a state equivalent to that in which a large amount of EGR is performed is obtained, so that the generation of NOx is sufficiently suppressed. This is also advantageous for improving emissions.

  Since the opening timing α1 of the gas introduction valve 31b at the time of control in the special operation mode is set to a timing later than the opening timing β1 of the gas outlet valve 32b as described above, the gas is discharged from the preceding cylinders 2A and 2D. The exhaust valve 32 and the gas which are closed in the vicinity of the exhaust top dead center of the succeeding cylinders 2B and 2C in a two-cylinder connection state where the burned gas is introduced into the succeeding cylinders 2B and 2C via the inter-cylinder gas passage 22 By shortening the overlap period OB with the introduction valve 31b, the burnt gas introduced into the succeeding cylinders 2B and 2C from the inter-cylinder gas passage 22 is blown through the exhaust port 12 to the exhaust passage 20 side. Is effectively prevented, and the combustibility in the succeeding cylinders 2B and 2C can be secured. In addition, by opening the gas outlet valve 32b provided at the upstream end of the inter-cylinder gas passage 22 at an early stage, the flow resistance of the burned gas led out from the preceding cylinders 2A and 2D into the inter-cylinder gas passage 22 is increased. It is possible to prevent the increase and make the burned gas flow smoothly.

  Further, as described above, the opening timing α1 of the gas introduction valve 31b during the control of the special operation mode is set to the exhaust top dead center TDC of the subsequent cylinders 2B and 2C or the vicinity thereof together with the closing timing γ2 of the exhaust valve 32. By setting, since the exhaust valve 32 and the gas introduction valve 31b are not substantially lifted at the piston top dead center, the piston top dead center positions of the succeeding cylinders 2B and 2C are relatively high. Accordingly, the geometric compression ratio of the succeeding cylinders 2B and 2C can be set to a higher value than the preceding cylinders 2A and 2D, and the succeeding cylinders 2B and 2C can be effectively compressed and self-ignited. Can do. In the succeeding cylinders 2B and 2C, homogeneous combustion is performed in both the special operation mode and the normal operation mode. Therefore, the bottom surface of the combustion chamber 4, that is, the upper surface of the piston 3 is formed flat so as to be suitable for this homogeneous combustion. This also makes it possible to further increase the geometric compression ratio of the succeeding cylinders 2B and 2C compared to the preceding cylinders 2A and 2D.

  As shown in the above embodiment, when the valve opening timing α1 of the gas introduction valve 31b at the time of control in the special operation mode is set at or near the exhaust top dead center TDC of the subsequent cylinders 2B and 2C, During the 2C exhaust stroke, the burnt gas can be more effectively blown out to the exhaust passage 20 due to the introduction of the burned gas from the inter-cylinder gas passage 22 into the succeeding cylinders 2B and 2C. Thus, the combustibility in the succeeding cylinders 2B and 2C can be secured, and the effective compression ratio of the succeeding cylinders 2B and 2C can be sufficiently increased to effectively compress and ignite the succeeding cylinders 2B and 2C. . In addition, the valve opening timing β1 of the gas outlet valve 32b is set to a timing earlier than the exhaust top dead center TDC of the subsequent cylinders 2B and 2C by a predetermined crank angle, so that the gas from the preceding cylinders 2A and 2D to the inter-cylinder gas passage 22 can be obtained. Thus, it is possible to effectively prevent the flow resistance of the burned gas derived from increasing, and to make the burned gas flow smoothly.

  Further, as shown in the above embodiment, the exhaust valve 32 for discharging the exhaust gas of the subsequent cylinders 2B and 2C to the exhaust passage 20 and the gas introduction valve 31b are disposed adjacent to each other in the subsequent cylinders 2B and 2C. In this case, the burned gas introduced into the succeeding cylinders 2B and 2C from the inter-cylinder gas passage 22 tends to blow through the exhaust port 12 to the exhaust passage 20 side. Gas provided at the downstream end of the inter-cylinder gas passage 22 in a two-cylinder connection state in which the burned gas discharged from the preceding cylinders 2A and 2D is introduced into the subsequent cylinders 2B and 2C through the inter-cylinder gas passage 22 The valve opening timing α1 of the introduction valve 31b is set to a timing later than the valve opening timing β1 of the gas outlet valve 32b provided at the upstream end of the inter-cylinder gas passage 22, and the exhaust gas provided in the succeeding cylinders 2B and 2C. Valve 32 and gas introduction valve 3 By shortening the overlap period OB with 1b, there is an advantage that the effect of preventing the burned gas introduced into the succeeding cylinders 2B and 2C from being blown into the exhaust passage 20 is remarkably obtained.

  Furthermore, in the above embodiment, the closing timing α2 of the gas introduction valve 31b during the control of the special operation mode is set to a timing later than the closing timing β2 of the gas outlet valve 32b, and the gas introduction valve 31b is relatively moved. Since it is configured to be in the late closing state, a large flow loss occurs when the burned gas discharged from the preceding cylinders 2A and 2D is introduced into the succeeding cylinders 2B and 2C through the inter-cylinder gas passage 22. Therefore, it is possible to improve the combustibility of the succeeding cylinders 2B and 2C by sufficiently ensuring the introduction amount of the burned gas to the succeeding cylinders 2B and 2C. In addition, by setting the gas outlet valve 32b to a relatively early closing state, the amount of internal EGR in the preceding cylinders 2A and 2D is prevented from increasing, and a sufficient amount of fresh air is introduced into the preceding cylinders 2A and 2D. can do. Therefore, by improving the combustibility in the preceding cylinders 2A and 2D and introducing high-temperature burned gas into the succeeding cylinders 2B and 2C, the compression self-ignitability of the succeeding cylinders 2B and 2C can be effectively improved. In addition, there is an advantage that an engine output corresponding to the operating state can be obtained.

  In particular, as shown in the above embodiment, when the valve closing timing β2 of the gas outlet valve 32b during the control in the special operation mode is set to the exhaust top dead center TDC of the preceding cylinders 2A and 2D or in the vicinity thereof, the preceding cylinder The generation of internal EGR due to the gas outlet valve 32b being maintained in the open state after the exhaust top dead center TDC of 2A, 2D is effectively suppressed, and the amount of fresh air introduced into the preceding cylinders 2A, 2D is sufficiently reduced. Can be secured. Moreover, the subsequent cylinders 2B and 2C are set by setting the valve closing timing α2 of the gas introduction valve 31b on the side of the subsequent cylinders 2B and 2C to a timing later than the exhaust top dead center TDC of the preceding cylinders 2A and 2D by a predetermined crank angle. The burned gas can be introduced more smoothly into the inside, and the combustibility of the succeeding cylinders 2B and 2C can be reliably improved.

  Furthermore, in the above embodiment, the overlap period OB between the exhaust valve 32 and the gas introduction valve 31b of the succeeding cylinders 2B and 2C at the time of control in the special operation mode is set to 5 ° CA or less in terms of the crank angle. In the two-cylinder connection state in which the burned gas discharged from 2D is introduced into the succeeding cylinders 2B and 2C via the inter-cylinder gas passage 22, the exhaust valves 32 and the gas introduction valves 31b of the succeeding cylinders 2B and 2C overlap. By setting the period of the state to be substantially 0, it is possible to suppress as much as possible that the burned gas introduced into the succeeding cylinders 2B and 2C is blown through to the exhaust passage 20 side.

  Moreover, when the succeeding cylinders 2B and 2C reach the exhaust top dead center TDC, the exhaust valve 32 and the gas are configured so that the lift amount of the exhaust valve 32 and the gas introduction valve 31b is substantially zero. While preventing interference between the introduction valve 31b and the piston 3, the top dead center position of the piston 3 can be set high. Therefore, the output performance of the engine can be improved by increasing the geometric compression ratio of the succeeding cylinders 2B and 2C.

  In the above-described embodiment, the gas outlet valve 32b and the gas introduction valve 31b provided at the upstream end and the downstream end of the inter-cylinder gas passage 22 are configured by poppet valves having the same diameter, and the gas introduction valve 31b. Because the set load of the valve spring that urges the valve in the closing direction is set to a value higher than the set load of the valve spring that urges the gas outlet valve 32b in the closing direction, the burned gas discharged from the preceding cylinders 2A and 2D Is connected to the succeeding cylinders 2B and 2C through the inter-cylinder gas passage 22, and the gas introduction valve 31b is opened before the set time according to the pressure of the burnt gas acting on the gas introduction valve 31b. It is possible to effectively prevent the valve state and prevent the burnt gas from being blown through due to the early introduction of the burned gas into the succeeding cylinder 31b. In addition, by setting the set load of the valve spring that biases the gas outlet valve 32b in the closing direction to a relatively low value, there is an advantage that the open / close response of the gas outlet valve 32b can be improved. .

  On the other hand, in the full load region B and the idle operation region C on the high load side or the high rotation side, as described above, the first exhaust valve 32a and the first intake valve 31a are activated and the gas introduction valve 32b gas is introduced. When the valve 31b is stopped, the flow paths of the fresh air and burned gas are substantially as shown in FIG. 9, and the intake ports 11 and 11a and the exhaust ports 12a and 12 of the cylinders 2A to 2D are connected. Independently, fresh air is introduced into the intake ports 11 and 11a of the respective cylinders 2A to 2D from the intake passage 15, and burned gas is discharged from the exhaust ports 12a and 12 of the respective cylinders 2A to 2D into the exhaust passage 20. The In the control state of the normal operation mode, the intake air amount and the fuel injection amount are controlled so as to become the stoichiometric air-fuel ratio or slightly richer, so that the output performance in the full load region B is ensured and the idling Combustion stability in the operation region C is maintained, and early activation of the three-way catalyst 24 is achieved.

  Instead of the above-described embodiment configured to execute the normal operation mode control in the idle operation region C shown in FIG. 4, the special operation mode control may be executed in the idle operation region C. In this case, in the idle operation region C, as shown in FIG. 10A, the valve closing timing α2 of the gas introduction valve 31b is set to about 24 ° CA after the exhaust top dead center TDC of the preceding cylinders 2A and 2D. In addition, the opening / closing timing of the intake valve 31 provided in the preceding cylinders 2A, 2D is advanced by a predetermined crank angle, for example, 36 ° CA, compared to the normal time indicated by the broken line, and is higher than the idle operation region C. In the partial load region A, as shown in FIG. 10B, the valve closing timing α of the gas introduction valve 31b is set to about 84 ° CA after the exhaust top dead center TDC of the two preceding cylinders 2A and 2D. It is desirable to switch the opening / closing timing of the gas introduction valve 31b by the switching mechanism so that the gas introduction valve 31b is in an extremely late closed state.

  That is, as shown in FIG. 11, the valve operating mechanism of the gas introduction valve 31b is provided with a switching mechanism 35a that switches each valve between an operation state and a stop state and switches the valve opening period of the gas introduction valve 31b. ing. The switching mechanism 35a is supported by the camshaft 34 disposed above the gas introduction valve 31b, the rocker shaft 55 disposed between the camshaft 34 and the gas introduction valve 31b, and the rocker shaft 55. The first to third rocker arms 56 to 58 are provided. The camshaft 34 includes a first cam 52 for stopping a valve having a circular outer peripheral surface, a second cam 53 for driving a valve having a protrusion (cam nose) having a large lift amount, and a small lift amount. A third cam 54 for driving the valve having the protrusion 54 is integrally formed (see FIG. 12).

  The first rocker arm 56 is disposed at a position corresponding to the first cam 52, and an abutting portion 60 that abuts on the upper end of the valve shaft of the gas introduction valve 31b is provided at the distal end thereof. On the other hand, the second and third rocker arms 57 and 58 are disposed on both sides of the first rocker arm 56 so as to sandwich the first rocker arm 56, and are separated from the first rocker arm 56 and are not shown in the drawing. Therefore, they are urged so as to be pressed against the second and third cams 53 and 54, respectively. The second and third rocker arms 57 and 58 are configured to be connectable to the first rocker arm 56. Plungers (not shown) provided on the second and third rocker arms 57 and 58 are driven by the hydraulic oil supplied from the first and second hydraulic oil supply / discharge passages, and the tip ends thereof are The first rocker arm 56 and one of the second rocker arm 57 or the third rocker arm 58 are integrally connected by being inserted into a connecting hole (not shown) formed in the first rocker arm 56. It is designed to swing and displace.

  Specifically, the supply and discharge of the hydraulic oil from the first and second hydraulic oil supply / discharge passages are controlled by the first and second control valves provided in the first and second hydraulic oil supply / discharge passages. When the first rocker arm 56 and the second rocker arm 57 are integrally connected, the driving force of the second rocker arm 57 driven by the second cam 53 is transmitted to the first rocker arm 56 and the gas introduction valve 31b. Is driven to open and close at the timing shown in FIG. Further, since the first rocker arm 56 and the third rocker arm 58 are integrally connected, the driving force of the third rocker arm 58 driven by the third cam 54 is transmitted to the first rocker arm 56 to introduce the gas. The valve 31b is driven to open and close at the timing shown in FIG. When the first rocker arm 56 is disconnected from the second and third rocker arms 57 and 58, the gas introduction valve 31a is maintained in the closed state.

  In addition, the valve mechanism of the exhaust valve 31 of the preceding cylinders 2A and 2D is provided with a second cam 53 and a third cam 54 that are arranged so as to have different phases. By selectively connecting the second rocker arm 57 or the third rocker arm 58 to the first rocker arm 56, the timing for quick opening and closing as shown in FIG. 10A and the preceding as shown in FIG. The cylinders 2A and 2D are controlled to be switched to the normal timing when the exhaust top dead center TDC or an open state at the vicinity thereof. Note that the cam indicated by the phantom line R in FIG. 12 is a drive cam for the gas outlet valve 32b.

  As described above, in the idle operation region C of the engine, as shown in FIG. 10A, the closing timing α2 of the gas introduction valve 31b is set earlier than the partial load region A, for example, in the preceding cylinders 2A and 2D. Since the predetermined crank angle (for example, 24 ° CA) after the exhaust top dead center TDC is set, the effective compression ratio of the succeeding cylinders 2B and 2C can be effectively increased. Therefore, by setting the closing timing of the exhaust valve 32 provided in the succeeding cylinders 2B and 2C and the opening timing α1 of the gas introduction valve 31b at or near the exhaust top dead center of the succeeding cylinders 2B and 2C. In combination with the fact that the compression ratio can be set higher, the compression self-ignitability of the succeeding cylinders 2B and 2C can be improved, and the temperature of the burned gas derived from the preceding cylinders 2A and 2D is relatively low. Also in C, the subsequent cylinders 2B and 2C can be effectively compressed and self-ignited.

  Moreover, as shown in the above-described embodiment, in the engine idle operation region C, the opening / closing timing of the intake valves 31 of the preceding cylinders 2A, 2D is advanced by a predetermined crank angle (for example, 36 ° CA) compared to the normal time. Thus, when the intake valve 31 is configured to be closed quickly, the effective compression ratio of the preceding cylinders 2A and 2D can be increased to increase the combustion temperature. By introducing a fuel gas, the compression self-ignition property can be effectively improved.

  On the other hand, in the partial load region A on the higher load side than the idle operation region C, as shown in FIG. 10B, the valve closing timing α2 of the gas introduction valve 31b is set to the exhaust top dead center TDC of the preceding cylinders 2A and 2D. By setting the gas inlet valve 31b to an extremely late closed state by setting to about 84 ° CA later, the effective compression ratio of the succeeding cylinders 2B and 2C is reduced, so that the air-fuel mixture is generated before the compression top dead center TDC. It is possible to effectively suppress the occurrence of knocking that spontaneously ignites, and to prevent a decrease in output and the generation of abnormal noise due to the occurrence of negative torque resulting from knocking. Therefore, it is possible to expand the operation area A in which combustion by compression ignition can be performed in the succeeding cylinders 2B and 2C to a higher load area.

  In the idle region C, ignition assist control for assisting compression self-ignition is performed by igniting the air-fuel mixture in the succeeding cylinders 2B and 2C in the vicinity of the top dead center before the compression top dead center TDC. Also good. In this case, the burned gas derived from the preceding cylinders 2A and 2D is introduced into the succeeding cylinders 2B and 2C, and the in-cylinder temperature is raised, and then the mixture is ignited, so that the subsequent time is reached. There is an advantage that the cylinders 2B and 2C can be surely compressed and self-ignited.

It is a schematic plan view of the whole engine provided with the control apparatus which concerns on embodiment of this invention. It is a schematic sectional drawing, such as an engine main body. It is explanatory drawing which shows embodiment of the control apparatus which concerns on this invention. It is explanatory drawing which shows an operation area | region. It is a figure which shows the exhaust stroke of each cylinder, an intake stroke, fuel injection timing, ignition timing, etc. FIG. It is explanatory drawing which shows the distribution channel of the fresh air and gas of special operation mode. It is explanatory drawing which shows the opening / closing timing of the valve in special operation mode. It is explanatory drawing which shows the lift characteristic of a valve. It is explanatory drawing which shows the distribution channel of the fresh air and gas of normal operation mode. It is explanatory drawing which shows another example of the opening / closing timing of the valve in special operation mode. It is a perspective view which shows the specific structure of the switching mechanism provided in the valve operating mechanism. It is explanatory drawing which shows the specific structure of the cam provided in the switching mechanism.

Explanation of symbols

1 Engine body 2A, 2D 1st and 4th cylinders (preceding cylinder)
2B, 2C 2nd and 3rd cylinders (following cylinders)
DESCRIPTION OF SYMBOLS 9 Fuel injection valve 15 Intake passage 20 Exhaust passage 22 Inter-cylinder gas passage 31b Gas introduction valve 32 Subsequent cylinder exhaust valve 32b Gas derivation valve 42 Valve stop mechanism control means (operation mode control means)
44 Combustion control means

Claims (7)

  A multi-cylinder spark ignition engine in which the combustion cycle of each cylinder is set to have a predetermined phase difference, and exhaust is performed between a pair of cylinders in which the exhaust stroke and the intake stroke coincide in a partial load region of the engine. The burned gas discharged from the preceding cylinder in the stroke is directly introduced into the subsequent cylinder in the intake stroke through the inter-cylinder gas passage, and the burned gas discharged from the subsequent cylinder is guided to the exhaust passage. While the cylinder is connected, combustion is performed with the air-fuel ratio of the preceding cylinder being a lean air-fuel ratio larger than the stoichiometric air-fuel ratio, and burned gas having a lean air-fuel ratio is introduced from the preceding cylinder to the succeeding cylinder and newly supplied. An operation mode control means for controlling a special operation mode for burning in the subsequent cylinder together with fuel is provided, and a gas outlet valve is provided at the upstream end of the inter-cylinder gas passage connected to the preceding cylinder. In both cases, a gas introduction valve is provided at the downstream end of the inter-cylinder gas passage connected to the succeeding cylinder, and the opening timing of the gas introduction valve at the time of control in the special operation mode is set higher than the opening timing of the gas outlet valve. A spark-ignition engine characterized by being set late.   Set the opening timing of the gas inlet valve in the special operation mode at or near the exhaust top dead center of the subsequent cylinder, and set the gas outlet valve opening timing earlier than the exhaust top dead center of the subsequent cylinder. 2. The spark ignition engine according to claim 1, wherein the spark ignition engine is set.   The spark ignition type engine according to claim 1 or 2, wherein an exhaust valve for exhausting exhaust gas of the succeeding cylinder to the exhaust passage and the gas introduction valve are disposed adjacent to the succeeding cylinder.   The spark ignition according to any one of claims 1 to 3, wherein the closing timing of the gas introduction valve at the time of control in the special operation mode is set to a timing later than the closing timing of the gas outlet valve. Expression engine.   Set the closing timing of the gas lead-out valve in the special operation mode at or near the exhaust top dead center of the preceding cylinder, and set the closing timing of the gas introduction valve to a timing later than the exhaust top dead center of the preceding cylinder. The spark ignition engine according to claim 4, wherein the spark ignition engine is set.   The overlap period between the exhaust valve of the subsequent cylinder and the gas introduction valve during the control in the special operation mode is set to 5 ° or less in crank angle, according to any one of claims 1 to 5. Spark ignition engine.   A set of valve springs that configure the gas outlet valve and the gas introduction valve provided at the upstream end and the downstream end of the inter-cylinder gas passage with poppet valves of the same diameter and urge the gas introduction valve in the closing direction. The spark ignition engine according to any one of claims 1 to 6, wherein the load is set to a value higher than a set load of a valve spring that urges the gas outlet valve in the closing direction.

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