JP3925379B2 - Control device for a spark ignition engine with a supercharger - Google Patents

Description

[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a control device for a spark ignition type engine, and more particularly to a device for controlling the combustion state of each cylinder in order to improve fuel consumption and emissions in a multi-cylinder engine.
[0002]
[Prior art]
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. Provided with a fuel injection valve that directly injects fuel into the room, and in the low rotation and low load range, etc., stratified combustion is performed by injecting fuel from the fuel injection valve in the compression stroke, thereby realizing super lean combustion Is known (see, for example, Patent Document 1).
[0003]
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) alone as an exhaust gas purification catalyst is sufficient for NOx during lean operation. Since the purification performance cannot be obtained, a lean NOx catalyst is provided that adsorbs NOx in an oxygen-excess atmosphere and removes and reduces NOx in an oxygen concentration-reduced atmosphere as shown in Patent Document 1 above. When such a 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 the above publication, additional fuel is injected during the expansion stroke in addition to the main combustion. As a result, the air-fuel ratio of the exhaust gas is enriched and CO is generated, thereby promoting NOx separation and reduction.
[0004]
[Patent Document 1]
Japanese Patent Laid-Open No. 10-274085
[0005]
[Problems to be solved by the invention]
The engine that performs the conventional lean operation as described above is disadvantageous in cost because the lean NOx catalyst is required to ensure the NOx purification performance during the lean operation. Further, in order to maintain the purification performance of the lean NOx catalyst, it is necessary to temporarily enrich the air-fuel ratio by removing additional NOx or supplying additional fuel for reduction when the NOx adsorption amount increases as described above. In addition, if the fuel used contains a large amount of sulfur, regenerative processing such as heating the catalyst and supplying reducing material is required to eliminate sulfur poisoning of the lean NOx catalyst. To do.
[0006]
Moreover, when the air-fuel ratio becomes leaner than a certain level, the combustion speed becomes too slow and combustion near the end does not contribute to work, so there is a limit to fuel efficiency improvement by leaning in stratified combustion.
[0007]
As another method for improving fuel efficiency, compression self-ignition has been studied. This compression self-ignition is similar to a diesel engine, with the combustion chamber at high dead temperature and high pressure at the top dead center of the compression stroke. Even if the air-fuel ratio is very lean or a large amount of EGR is introduced, if such compression self-ignition is performed, the entire combustion chamber burns at once, so it does not contribute to work. Combustion is avoided, which is advantageous for improving fuel efficiency. However, in a normal spark ignition engine (gasoline engine), forced ignition is required for combustion, and in order to perform compression self-ignition, special measures are taken to significantly increase the temperature or pressure in the combustion chamber. Therefore, it has been difficult to increase the temperature or pressure in the combustion chamber to such an extent that compression self-ignition occurs in a partial load region where fuel consumption improvement is required, while avoiding knocking in a high load region.
[0008]
Therefore, the present applicant is in the exhaust stroke between a pair of cylinders in which the exhaust stroke and the intake stroke overlap in a partial load region of the engine in order to have a significant fuel efficiency improvement effect by using both lean combustion and compression self-ignition. The burned gas discharged from the preceding cylinder is in a two-cylinder connection state in which the burned gas discharged from the preceding cylinder is directly introduced into the succeeding cylinder in the intake stroke via the inter-cylinder gas passage. A spark ignition type engine in which combustion is performed by forced ignition at a fuel ratio and fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinder in the subsequent cylinder and combustion is performed by compression self-ignition. A technology related to the control device has been filed (Japanese Patent Application No. 2002-29836).
[0009]
Based on such a technique, the present invention is capable of effectively performing combustion by compression self-ignition in a subsequent cylinder in a wider area, and can improve the fuel efficiency and emission improvement effect. An engine control device is provided.
[0010]
[Means for Solving the Problems]
The invention according to claim 1 is a multi-cylinder spark ignition type 4-cycle engine in which the combustion cycle of each cylinder is performed with a predetermined phase difference, and the intake / exhaust and combustion of the engine in a partial load region of the engine. The control mode for the state is a special operation mode, and in this special operation mode, the burned gas discharged from the preceding cylinder in the exhaust stroke between the pair of cylinders in which the exhaust stroke and the intake stroke overlap is directly in the intake cylinder In the preceding cylinder, the air-fuel ratio is higher than the stoichiometric air-fuel ratio while the two-cylinder connection state is established such that the gas discharged from the subsequent cylinder is introduced into the exhaust passage and led to the exhaust passage. A spark ignition type air that causes combustion to be performed and fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinder to cause combustion in the subsequent cylinder. A control device for a gin, wherein, in at least a part of the operation region in the special operation mode, in the preceding cylinder, the exhaust valve is closed before the exhaust stroke top dead center for a predetermined period of time. Control is performed so that part of the burned gas remains in the cylinder and the valve opening period of the intake valve overlaps the valve opening period of the exhaust valve, and combustion is performed by compression self-ignition in the subsequent cylinder. While being controlled, a supercharger for assisting the intake air in the preceding cylinder is provided.
[0011]
According to the present invention, when combustion is performed by compression self-ignition in the subsequent cylinder in the special operation mode, the preceding cylinder can achieve the fuel efficiency improvement effect by improving the thermal efficiency and the pumping loss by the lean combustion, and the subsequent cylinder can compress the self-compression. Improvement of fuel efficiency can be obtained by improving combustion efficiency by ignition and reducing pumping loss. Further, the NOx generation amount is suppressed to a relatively low level by performing combustion at a lean air-fuel ratio in the preceding cylinder, and a large amount of EGR (exhaust gas recirculation) is introduced in the succeeding cylinder by introducing burned gas from the preceding cylinder. Since it becomes the state equivalent to what is performed, generation | occurrence | production of NOx is fully suppressed and it contributes to purification | cleaning of waste gas. Furthermore, since the exhaust gas from the subsequent cylinders to the exhaust passage can have a stoichiometric air-fuel ratio, it is possible to sufficiently purify exhaust gas using only a three-way catalyst, and a relatively expensive lean NOx catalyst is required. This will also lead to cost reduction.
[0012]
Further, since fresh air is taken in and burned with the burned gas remaining in the preceding cylinder, the action of suppressing knocking in the preceding cylinder is enhanced by increasing the burned gas component. Also, since the preceding cylinder takes in fresh air into the relatively high-temperature residual burned gas and burns it, the burned gas becomes even hotter, and this high-temperature burned gas is introduced into the succeeding cylinder. While the compression self-ignition in the cylinder is reliably and smoothly performed, the action of suppressing knocking in the subsequent cylinder is enhanced by increasing the burned gas component corresponding to EGR in the gas. Furthermore, there is a concern about an increase in pumping loss accompanying the early closing of the exhaust valve of the preceding cylinder. However, since the intake valve overlaps the exhaust valve, an increase in pumping loss can be avoided. it can.
[0013]
Here, in the preceding cylinder, some of the burned gas remains, and the intake valve is opened in the exhaust stroke in order to overlap the exhaust valve, so the output decreases due to insufficient intake of fresh air However, according to the present invention, since the supercharger is provided to assist the intake of the preceding cylinder, it is possible to force fresh air into the preceding cylinder by this supercharger. Therefore, it is possible to prevent a decrease in output due to a decrease in intake air amount.
[0014]
In the present invention, it is preferable that the closing timing of the intake valve in the preceding cylinder is set to substantially coincide with the intake stroke bottom dead center (claim 2). In this way, the effective compression ratio of the engine is improved and the in-cylinder temperature can be maintained at a higher temperature, so that even higher temperature burned gas can be introduced into the subsequent cylinder. Therefore, for example, even in a low load range where the in-cylinder temperature does not easily rise, it is possible to more reliably ensure the self-ignition performance in the subsequent cylinder and to perform the combustion by the compression self-ignition in the subsequent cylinder. The area can be expanded, and further improvement in fuel consumption and exhaust gas purification can be promoted.
[0015]
Further, in the present invention, a variable valve timing mechanism for changing the valve timing of the intake and exhaust valves in the preceding cylinder is provided, and in a higher load region than the partial operation region in the operation region to be the special operation mode, The variable valve timing mechanism is set so that the exhaust valve of the preceding cylinder is closed after the exhaust stroke top dead center has elapsed for a predetermined period, and the intake valve of the preceding cylinder has passed the intake stroke for a predetermined period. It is preferable that the valve is set to be closed (Claim 3). If it does in this way, the optimal valve timing according to the driving | running state of an engine can be set with a variable valve timing mechanism, and the output performance according to the driving | running state of an engine can be ensured. In addition, for example, knocking in the succeeding cylinder, which may be caused when the burned gas introduced into the succeeding cylinder becomes excessively high in temperature, can be effectively prevented. In other words, in the high load range, both the preceding and succeeding cylinders have high internal temperatures. In this state, the preceding cylinder takes in fresh air into the burned gas and burns it. When burnt gas is introduced into the succeeding cylinder, the in-cylinder temperature of the preceding and succeeding cylinders becomes higher than necessary, and abnormal combustion such as knocking may occur. Therefore, when configured as described above, the temperature rise due to the remaining burned gas in the preceding cylinder is suppressed, and the burned gas temperature in the preceding cylinder introduced to the succeeding cylinder is lowered to effectively knock the preceding and succeeding cylinders. In addition, since no burned gas remains in the preceding cylinder, fresh air can be taken in sufficiently, and the output performance of the engine can be sufficiently exhibited.
[0016]
Further, in the control device provided with the variable valve timing mechanism, the intake / exhaust valves in the preceding cylinder are made to correspond to the intake / exhaust strokes in a higher load range than the operation range in which the special operation mode is set. The cylinder is set to switch to a normal operation mode in which each cylinder is independently burned by forced ignition. In this normal operation mode, a part of the exhaust gas in each cylinder is externally provided with an EGR cooler. It is preferable to introduce into the intake passage via the EGR passage. In this way, the engine output performance corresponding to the load can be secured, and part of the exhaust gas cooled by the EGR cooler is introduced into the intake passage via the external EGR, thereby increasing the temperature in the cylinder. Can be effectively prevented.
[0017]
Moreover, the said supercharger is although it does not specifically limit, It is preferable to use a turbocharger as a supercharger. As described above, when the turbocharger is a turbocharger, the exhaust pressure becomes high, so it is easy for the burned gas to remain in the preceding cylinder, and the burned gas introduced into the succeeding cylinder is further heated. And the self-ignitability of the succeeding cylinder can be improved.
[0023]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings.
[0024]
FIG. 1 shows a schematic configuration of an engine according to an embodiment of the present invention, 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.
[0025]
A spark plug 7 is provided at the top of the combustion chamber 4 of each cylinder 2, 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.
[0026]
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 and opened for a time corresponding to the pulse width at the pulse input timing. 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) through 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. Is configured.
[0027]
Further, intake ports 11, 11a, 11b and exhaust ports 12, 12a, 12b are opened to the combustion chambers 4 of the respective cylinders 2A to 2D, and an intake passage 15 and an exhaust passage 20 are connected to these ports. Each port is opened and closed by intake valves 31, 31a, 31b and exhaust valves 32, 32a, 32b.
[0028]
Each cylinder performs a cycle consisting of intake, compression, expansion, and exhaust strokes with a predetermined phase difference. In the case of a four-cylinder engine, the first cylinder 2A, second cylinder from one end in the cylinder row direction When the cylinder 2B, the third cylinder 2C, and the fourth cylinder 2D are called, as shown in FIGS. 6 and 7, the above cycle is performed in the order of the first cylinder 2A, the third cylinder 2C, the fourth cylinder 2D, and the second cylinder 2B. It is performed with a phase difference of about 180 ° in angle. 6 and 7, EX is an exhaust stroke, IN is an intake stroke, F is fuel injection, S is forced ignition, and a star mark in the drawings indicates that compression self-ignition is performed. ing.
[0029]
Between a pair of cylinders in which the exhaust stroke and the intake stroke are substantially overlapped, a cylinder on the intake stroke side (this cylinder is referred to as a preceding cylinder in this specification) when the exhaust stroke and the intake stroke are substantially overlapped (this cylinder is referred to as a preceding cylinder). The inter-cylinder gas passage 22 is provided so that the burned gas can be guided as it is to the subsequent cylinder in the specification). In the four-cylinder engine of this embodiment, as shown in FIGS. 6 and 7, the exhaust stroke (EX) of the first cylinder 2A and the intake stroke (IN) of the second cylinder 2B substantially overlap, and the fourth cylinder 2D Since the exhaust stroke (EX) and the intake stroke (IN) of the third cylinder 2C substantially overlap, the first cylinder 2A and the second cylinder 2B, and the fourth cylinder 2D and the third cylinder 2C form a pair, respectively. The cylinder 2A and the fourth cylinder 2D are the preceding cylinders, the second cylinder 2B and the third cylinder 2C are the succeeding cylinders.
[0030]
The intake / exhaust port of each cylinder and the intake passage, exhaust passage, and inter-cylinder gas passage connected to the cylinder are specifically configured as follows.
[0031]
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 20. And a second exhaust port 12b for leading the burned gas to the succeeding cylinder via the inter-cylinder gas passage 22. Further, the second cylinder 2B and the third cylinder 2C, which are the subsequent cylinders, respectively, receive the first intake port 11a for introducing fresh air and the burned gas from the preceding cylinder through the inter-cylinder gas passage 22. A second intake port 11b for introduction and an exhaust port 32 for sending burned gas to the exhaust passage 20 are provided.
[0032]
In the example shown in FIG. 1, the intake ports 11 in the first and fourth cylinders 2A and 2D and the first intake ports 11a in the second and third cylinders 2B and 2C are two on the left side of the combustion chamber 4 for each cylinder. The first exhaust port 12a and the second exhaust port 12b in the first and fourth cylinders 2A and 2D and the second intake port 11b and the exhaust in the second and third cylinders 2B and 2C are provided in parallel on the half side. The port 12 is provided in parallel on the right half side of the combustion chamber 4.
[0033]
The intake port 11 in the first and fourth cylinders 2A and 2D and the first intake port 11a in the second and third cylinders 2B and 2C are connected to the downstream ends of the branch intake passages 16 for each cylinder in the intake passage 15. Yes. 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.
[0034]
An upstream end of a branch exhaust passage 21 for each cylinder in the exhaust passage 20 is connected to the first exhaust port 12a in the first and fourth cylinders 2A and 2D and the exhaust port 12 in the second and third cylinders 2B and 2C. Yes. 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, and the first, fourth cylinder 2A, The upstream end of the inter-cylinder gas passage 22 is connected to the 2D second exhaust port 12b, and the downstream end of the inter-cylinder gas passage 22 is connected to the second intake port 11b of the second and third cylinders 2B and 2C as the subsequent cylinders. Is connected.
[0035]
The inter-cylinder gas passage 22 has a linear O in which output varies linearly according to oxygen concentration. ₂ A sensor 25 is provided, and the fuel injection amount for the preceding cylinders 2A and 2D having a predetermined lean air-fuel ratio is feedback-controlled according to the output thereof. In the apparatus of the present embodiment, the inter-cylinder gas passage 22 is a relatively short passage that connects adjacent cylinders, and the gas discharged from the preceding cylinders 2A and 2D passes through the passage 22. The heat radiation is kept relatively small.
[0036]
Further, a supercharger that assists intake air is provided for the preceding cylinders 2A and 2D, and in this embodiment, a turbocharger 27 is provided. In the present embodiment, the supercharger 27 is driven so as to assist the intake of the preceding cylinders 2A and 2D in a special operation mode and a normal operation mode, which will be described later. May be limited. The supercharger 27 is connected to the turbine 28 provided in the exhaust passage 20 communicating with the exhaust ports 12, 12a, 12b via the branch exhaust passage 21, and the intake passage 15 communicating with the intake ports 11, 11a, 11b. A compressor 29 provided in conjunction with the turbine 28, and the turbine 28 is rotated by the energy of the exhaust gas flowing through the exhaust passage 20, and the compressor 29 is also rotated in conjunction with the rotation of the turbine 28. The intake air is supercharged by the rotation of. The turbocharger is not limited to the turbocharger 27 described above, and may be a mechanical supercharger such as a supercharger. Since the pressure becomes higher, it becomes easier for the burned gas to remain in the preceding cylinders 2A and 2D as will be described later, and the burned gas introduced into the succeeding cylinders 2B and 2C can be further increased in temperature. This is advantageous in that the self-ignitability in 2C can be improved. The supercharging pressure by the supercharger 27 is sufficient to assist the supply of fresh air to the preceding cylinders 2A and 2D in which the burned gas remains and the intake pressure decreases as will be described later. It ’s enough.
[0037]
Further, an external EGR passage 26 for introducing a part of the exhaust gas in each cylinder into the intake passage 15 is provided downstream of the exhaust passage 20, and an EGR cooler that cools the exhaust gas is provided in the external EGR passage 26. An external EGR control valve 26b that opens and closes 26a and the external EGR passage 26 is provided. The exhaust gas cooled by the EGR cooler 26a is introduced into the intake passage 15 with the opening and closing of the external EGR control valve 26b. The external EGR passage 26 is used to prevent the occurrence of knocking by suppressing an increase in the in-cylinder temperature. In the present embodiment, the external EGR passage 26 is controlled to function in a normal operation mode described later.
[0038]
In the exhaust passage 20, downstream of the branch exhaust passage 21, an air-fuel ratio is detected by detecting the oxygen concentration in the exhaust gas. ₂ A sensor 23 is provided. This O ₂ The sensor 23 outputs λO whose output changes suddenly near the theoretical air-fuel ratio. ₂ This is sensor 23. ₂ Based on the output of the sensor 23, the fuel injection amount for the subsequent mechanisms 2B and 2C (including the cylinders 2A and 2D when each cylinder is in an independent state) is feedback-controlled. Furthermore O ₂ A three-way catalyst 24 is provided in the exhaust passage 21 downstream of the sensor 23 for exhaust purification. As is generally known, the three-way catalyst 24 is highly purified against HC, CO, and NOx when the air-fuel ratio of the exhaust gas is close to the stoichiometric air-fuel ratio (that is, the excess air ratio λ is λ = 1). It is a catalyst showing performance.
[0039]
The intake / exhaust valves for opening and closing the intake / exhaust ports of each cylinder and the valve operating mechanism for these valves are as follows.
[0040]
The intake port 11, the first exhaust port 12a, and the second exhaust port 12b in the first and fourth cylinders 2A, 2D are respectively provided with an intake valve 31, a first exhaust valve 32a, and a second exhaust valve 32b. A first intake valve 31a, a second intake valve 31b, and an exhaust valve 32 are provided in the first intake port 11a, the second intake port 11b, and the exhaust port 12 in the No. 3 cylinders 2B and 2C, respectively.
[0041]
Each of these valves has a predetermined timing by a valve mechanism according to the intake / exhaust stroke of each corresponding cylinder, that is, according to the vertical movement of each piston 3 (movement between top dead center and bottom dead center). However, the opening and closing timing does not always coincide with the top dead center or bottom dead center of the piston 3, and the crank angle is advanced or retarded by several degrees to several tens of degrees as necessary. You may set it at the time of letting. In particular, in the present embodiment, the valve timing is changed according to the operating state of the engine or the valve can be kept in a fully closed state by a variable valve timing mechanism provided in the valve operating mechanism.
[0042]
Specifically, in this embodiment, cam switching mechanisms 51 and 61 that change cams according to operating conditions are adopted as the variable valve timing mechanism, and intake valves of the first and fourth cylinders 2A and 2D, which are the preceding cylinders, are adopted. The valve 31 and the second exhaust valve 32b can change the valve timing over three stages, while the other valves can change the valve timing over two stages.
[0043]
FIG. 4 is a partial perspective view showing a cam switching mechanism 51 that can be switched in three stages. As described above, the cam switching mechanism 51a and the intake valve 31 provided for the second exhaust valve 32b indicated by a two-dot chain line. And a cam switching mechanism 51b provided for the purpose.
[0044]
The cam switching mechanism 51a provided for the second exhaust valve 32b includes a camshaft 33 provided above the second exhaust valve 32b and provided with a plurality of types of cams according to the switching stage, and the camshaft 33. And a second exhaust valve 32b, a rocker arm set 56 supported by a rocker shaft 55 is provided.
[0045]
The camshaft 33 is provided with a plurality of types of cams corresponding to the switching stage, that is, first to third types of cams 52, 53, 54 adjacent to each other in the axial direction, and these cams are integrated with the camshaft 33. It is supposed to rotate. Each cam has a different lift characteristic, that is, the valve opening / closing valve timing, valve opening period, lift amount and the like are determined by each cam. The shape and operation of each cam will be described later.
[0046]
The rocker arm set 56 is an assembly of three types of rocker arms, first to third rocker arms 57, 58, and 59, and transmits and opens the cam driving force to the second exhaust valve 32b.
[0047]
Specifically, the first rocker arm 57 is provided with a valve contact portion 60 at the tip thereof, and the valve contact portion 60 contacts the upper end of the valve shaft of the second exhaust valve 32b at an appropriate position. On the other hand, the second and third rocker arms 58 and 59 are provided on both sides of the first rocker arm 57 so as to be separated from the first rocker arm 57, and the second and third cams 53 and 54 are respectively provided by springs not shown. It is supposed to be pressed. Therefore, when the rocker arms of the rocker arm set 56 are independently movable as shown in the figure, the upper surfaces of the rocker arms 57, 58, 59 are in contact with the outer peripheral portions of the first to third cams 52, 53, 54. In contact with each other, the rocker shaft 55 is pivoted up and down according to the shape of the cam contact portion (the rotation radius of each cam).
[0048]
Further, the first to third rocker arms 57, 58, 59 are individually provided as described above, but each inside a plurality of plunger holes (not shown) communicated between the rocker arms. The first rocker arm 57 and the second rocker arm 58 or the first rocker arm 57 and the third rocker arm 59 can be connected by a plurality of plungers (not shown) that slide by hydraulic operation. The integrated rocker arms are linked. These plungers are slid by the oil flowing through the first and second hydraulic oil supply / discharge passages 36, 38, while the oil is provided in the respective hydraulic oil supply / discharge passages 36, 38. The flow is controlled by the first and second control valves 37 and 39 (see FIG. 3). Accordingly, the cam switching mechanism 51 controls the first and second control valves 37 and 39 to slide the plungers, and the movement of these plungers causes the rocker arm separation state, the first and the first to move. The two rocker arms 57 and 58 can be switched to an integrated state, and the first and third rocker arms 57 and 59 can be switched to three states.
[0049]
The operation of the cam switching mechanism 51 will be described together with the shape of each cam.
[0050]
The first cam 52 is a valve stopping cam and has an outer peripheral shape concentric with the camshaft 33. Therefore, when the first and second control valves 37 and 39 are controlled to separate the rocker arms, that is, when the first cam 52 is selected by the cam switching mechanism 51, the upper surface of the first rocker arm 57 is The cam 52 is always in contact with the outer peripheral surface of the cam 52 and does not swing even if the cam shaft 33 rotates, and the second exhaust valve 32b in contact with the valve contact portion 60 of the first rocker arm 57 does not open or close. .
[0051]
The second cam 53 is a low-load cam, and includes a portion having the same outer peripheral shape as the first cam 52 and a portion having an outer peripheral shape protruding radially outward therefrom. The protruding portion of the second cam 53 is adjusted so that the second exhaust valve 32b is closed a predetermined period before the exhaust stroke top dead center. Therefore, as the camshaft 33 rotates, the second rocker arm 58 swings downward by a predetermined amount at a predetermined crank angle according to the protruding portion of the second cam 53, while the second rocker arm 58 is swung downward by the first rocker arm 57. When the second cam 53 is selected by the cam switching mechanism 51, the first rocker arm 57 swings as the second rocker arm 58 swings, and the second exhaust valve 32b opens for a predetermined amount during a predetermined period, and closes a predetermined period before the top dead center of the exhaust stroke.
[0052]
The third cam 54 is a high-load cam, and is set to function in the same manner as a general cam of a conventional engine. The third cam 54 has a portion having the same outer peripheral shape as the first cam 52, and a radial direction therefrom. And a portion having an outer peripheral shape protruding outward. This protruding portion is set at an angle slightly different from the protruding portion of the second cam 53, and is adjusted so that the second exhaust valve 32b is closed after a predetermined period of time has elapsed from the top dead center of the exhaust stroke. Therefore, as the camshaft 33 rotates, the third rocker arm 59 swings downward by a predetermined amount at a predetermined crank angle according to the protruding portion of the third cam 54, while the third rocker arm 59 is moved to the first rocker arm 57. When the third cam 54 is selected by the cam switching mechanism 51, the first rocker arm 57 swings with the swing of the third rocker arm 59, and the second exhaust valve. The valve 32b is opened by a predetermined amount during a predetermined period, and is closed after a predetermined period from the exhaust stroke top dead center.
[0053]
That is, in the present embodiment, the third cam 54 is driven so that the second exhaust valve 32b is driven at a general opening / closing valve timing, valve opening period, etc. corresponding to the intake / exhaust stroke, similarly to the conventional engine. On the other hand, the second cam 53 is closed before the top dead center of the exhaust stroke by advancing the second exhaust valve 32b as a whole as compared with the second exhaust valve 32b driven by the third cam 54. It is set to speak. The first to third cams 52, 53, 54 are selected by the cam switching mechanism 51 according to the operating state of the engine, and the second exhaust valve 32b is set to be driven according to the outer peripheral shape of each cam. Yes.
[0054]
On the other hand, the three-stage switchable cam switching mechanism 51b provided for the intake valve 31 is configured similarly to the three-stage switchable cam switching mechanism 51a provided for the second exhaust valve 32b. However, the opening timing of the intake valve 31 when the second cam 53 is employed is set to overlap the second exhaust valve 32b, and the closing timing is set to coincide with the bottom dead center of the intake stroke. Is different. On the other hand, as with the cam switching mechanism 51 provided for the second exhaust valve 32b, the third cam 54 is driven at a general opening / closing valve timing of a conventional engine, that is, an intake valve. 31 is set so as to correspond to the intake stroke so as to be opened a predetermined period before the intake stroke top dead center, and set so that the valve is closed after a predetermined period has elapsed from the intake stroke bottom dead center. Has been.
[0055]
Further, for the valves other than the intake valve 31 and the second exhaust valve 32b of the first and fourth cylinders 2A and 2D which are the preceding cylinders, the valve timing can be changed by a cam switching mechanism 61 which can be switched in two stages. However, the configuration is the same as that of the cam switching mechanism 51 except for some configurations such as few plunger holes, and each valve is stopped by the first cam 62 and the second cam 63. And switching to the operating state. Therefore, the description of the cam switching mechanism 61 is omitted. The valve that is activated by the two-stage switchable cam switching mechanism 61 is set to a general valve timing of a conventional engine.
[0056]
In this embodiment, a variable valve timing / lift mechanism (VVL) that changes the cam according to the operating state is adopted as the variable valve timing mechanism. However, the present invention is not limited to this. A known valve timing variable mechanism such as a variable valve timing system (VTC) for advancing the entire period may be employed. In this case, a publicly known valve stop mechanism, a control means for controlling the same, and the like are provided so that a special operation mode and a normal operation mode as described later can be switched.
[0057]
FIG. 3 shows the configuration of the drive and control system. In this figure, an ECU (control unit) 40 for engine control composed of a microcomputer or the like is provided with an air flow sensor 19 and an O ₂ Sensor 23 and linear O ₂ A signal from the sensor 25 is input, and further, a signal from an engine speed sensor 47 for detecting the engine speed in order to determine an operating state, an accelerator position sensor 48 for detecting an accelerator position (accelerator pedal depression amount), and the like. Have been entered. Further, from the ECU 40 to the ignition circuit 8, each fuel injection valve 9, the actuator 18 of the multiple throttle valve 17, the first and second control valves 37 and 39, and the external EGR control valve 26b. A control signal is output.
[0058]
The ECU 40 includes an operation state determination unit 41, a cam switching control unit 42, an intake air amount control unit 43, a combustion control unit 44, and an external EGR control unit 49.
[0059]
As shown in FIG. 5, the operating state discriminating means 41 is a control in which the operating range of the engine is divided into an operating range A (partial load range) on the low load low rotation side and an operating range B on the high load side or high rotation side. And a control map which is divided into an operation region A1 on the low and medium load side and an operation region A2 on the high load side in the operation region A, and the rotational speed sensor 47 and the accelerator opening sensor 48 are provided. It is determined whether the operating state of the engine (engine speed and engine load), which is examined by a signal from the above, is in the above-mentioned operating regions A, B and operating regions A1, A2. Based on this determination, in the operation region A on the low load low rotation side, a special operation mode is selected in which the burned gas discharged from the preceding cylinder in the exhaust stroke is directly introduced into the subsequent cylinder in the intake stroke and burned. In the high load side or high rotation side operation region B, the normal operation mode in which each cylinder is burned independently is selected.
[0060]
Furthermore, when the operation state determination unit 41 is in the operation region A in which the special operation mode is selected, the operation state determination unit 41 determines whether the region A is in the low / medium load region A1 or the high load region A2. The optimum cam is selected by the cam switching control means 42 based on the determination result, and the fuel injection timing, the ignition method, and the like are selected by the combustion control means 44.
[0061]
The cam switching control means 42 controls the first and second control valves 37 and 39 on the basis of the output of the operating state determining means 41, that is, according to the operating state of the engine. To control.
[0062]
Operation area A1 (low / medium load area in special operation mode)
First exhaust valve 32a and first intake valve 31a
The first cam 62 is selected by the cam switching mechanism 61 and is stopped.
Intake valve 31 and second exhaust valve 32b
The second cam 53 is selected by the cam switching mechanism 51 and is activated.
The second intake valve 31b and the exhaust valve 32
The second cam 63 is selected by the cam switching mechanism 61 and is in an operating state.
Operation area A2 (High load area in special operation mode)
First exhaust valve 32a and first intake valve 31a
The first cam 62 is selected by the cam switching mechanism 61 and is stopped.
Intake valve 31 and second exhaust valve 32b
The third cam 54 is selected by the cam switching mechanism 51 and is in an operating state.
The second intake valve 31b and the exhaust valve 32
The second cam 63 is selected by the cam switching mechanism 61 and is in an operating state.
Operation area B (normal operation mode)
First exhaust valve 32a and first intake valve 31a
The second cam 63 is selected by the cam switching mechanism 61 and is in an operating state.
・ Intake valve 31
The third cam 54 is selected by the cam switching mechanism 51 and is in an operating state.
・ Second exhaust valve 32b
The first cam 52 is selected by the cam switching mechanism 51 and stopped.
・ Second intake valve 31b
The first cam 62 is selected by the cam switching mechanism 61 and is stopped.
・ Exhaust valve 32
The second cam 63 is selected by the cam switching mechanism 61 and is in an operating state.
[0063]
The intake air amount control means 43 controls the opening degree of the throttle valve 17 (throttle opening degree) by controlling the actuator 18, and obtains a target intake air amount from a map or the like according to the operating state. The throttle opening is controlled according to the target intake air amount. In this case, in the operation region A that is set to the special operation mode, in the second and third cylinders 2B and 2C that are the subsequent cylinders, the first cylinder that is the preceding cylinder in a state where the intake introduction from the branch intake passage 16 is blocked. Combustion is performed while the ratio of excess air in the gas introduced from the fourth cylinders 2A and 2D to the newly supplied fuel is made to be a lean air-fuel ratio, so the combustion is performed by the supercharger 27 and the preceding cylinder 2A , 2D is an amount of air necessary for the combustion of fuel corresponding to the required torque for the preceding and succeeding two cylinders (the amount of air that is the stoichiometric air-fuel ratio with respect to the amount of fuel for the two cylinders) ), The throttle opening is adjusted.
[0064]
The combustion control means 44 includes a fuel injection control means 45 and an ignition control means 46. The fuel injection control means 45 causes the fuel injection amount and injection from the fuel injection valves 9 provided in the respective cylinders 2A to 2D. The timing is controlled in accordance with the operating state of the engine, and the ignition control means 46 controls the ignition timing and the ignition stop in accordance with the operating state. In particular, when the operating state is in the operating region A and in the operating region B in FIG. 5, and even in the operating region A, the combustion state is in the operating region A1 and in the operating region A2. (Control of fuel injection and control of ignition) are changed.
[0065]
That is, when the operation state is in the operation region A on the low load and low rotation side, the air-fuel ratio is set to the stoichiometric air-fuel ratio for the preceding cylinders (first and fourth cylinders 2A and 2D) as control in the special operation mode. The fuel injection amount is controlled so that the lean air-fuel ratio is larger than that, the injection timing is set so that the fuel is injected in the compression stroke and the mixture is stratified, and near the compression top dead center. The ignition timing is set so that forced ignition is performed. On the other hand, for the subsequent cylinders (second and third 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, so that the theoretical air-fuel ratio is substantially increased. The fuel injection amount is controlled so as to be, and when in the operation region A (including the operation region A1 and the operation region A2), the injection timing is set so that the fuel is injected in the intake stroke, and the compression self In order to perform ignition, it is set to stop forced ignition.
[0066]
Further, when the operation state is in the operation region B on the high load side or the high rotation side, the control is performed in the normal operation mode so that the air-fuel ratio of each cylinder 2A to 2D is the stoichiometric air-fuel ratio or less. The injection amount is controlled so that, for example, the stoichiometric air-fuel ratio is set in most of the operating range B, and is made richer than the stoichiometric air-fuel ratio in the fully open load and the operating range in the vicinity thereof. In this case, the injection timing is set so that the air-fuel mixture is made uniform by injecting fuel to each of the cylinders 2A to 2D and the cylinders 2A to 2D are forcedly ignited. To.
[0067]
The external EGR control means 49 is in a closed state when the normal operation mode is selected based on the result of the operation state determination means 41, that is, when the engine operation state is in the operation region B. The external EGR control is performed so that the external EGR control valve 26b is opened, and external EGR for introducing the partial exhaust gas cooled by the EGR cooler 26a into the intake passage 15 is executed.
[0068]
The operation of the apparatus of the present embodiment as described above will be described with reference to FIGS.
[0069]
In the operation region A, the special operation mode is set. As described above, the first exhaust valve 32a and the first intake valve 31a are stopped by the cam switching mechanism 61, and the second exhaust valve 32b and the second intake valve 31b are stopped by the cam switching mechanism 51. 9 is in the operating state and the cam switching mechanism, the actual flow path of fresh air and gas is as shown in FIG. 9, and the exhaust gas discharged from the preceding cylinders (first and fourth cylinders) 2A and 2D is changed. The fuel gas is introduced as it is into the subsequent cylinders (second and third cylinders) 2B and 2C through the inter-cylinder gas passage 22, and only the gas discharged from the subsequent cylinders 2B and 2C is guided to the exhaust passage 20. Such a two-cylinder connection state is established.
[0070]
In this state, fresh air is introduced into the preceding cylinders 2A and 2D while being assisted by the supercharger 27 from the intake passage 15 in the intake stroke (arrow a in FIG. 9), and in the preceding cylinders 2A and 2D, linear O ₂ Fuel is injected in the compression stroke while the fuel injection amount is feedback controlled so that the air-fuel ratio detected by the sensor 25 becomes a lean air-fuel ratio larger than the stoichiometric air-fuel ratio, and ignition is performed at a predetermined ignition timing, Stratified combustion is performed at a lean air-fuel ratio (see FIGS. 6 and 7).
[0071]
On the other hand, burned gas discharged from the preceding cylinders 2A and 2D is introduced into the succeeding cylinders 2B and 2C through the gas passage 22 during a period in which the intake strokes of the preceding cylinders 2A and 2D overlap with the exhaust strokes of the succeeding cylinders 2B and 2C. (The white arrow in FIGS. 6 and 7 and the arrow b in FIG. 9). In the succeeding cylinders 2B and 2C, the fuel is supplied to the burned gas having the lean air-fuel ratio introduced from the preceding cylinders 2A and 2D, and the fuel injection amount is controlled so as to become the stoichiometric air-fuel ratio. In (operating region A1 and operating region A2), after fuel is injected in the intake stroke, compression self-ignition is performed due to an increase in pressure and temperature in the combustion chamber near the top dead center of the compression stroke (FIGS. 6 and 7). reference).
[0072]
In this way, in the preceding cylinders 2A and 2D, the thermal efficiency is increased by stratified combustion in lean, and the pumping loss is reduced by reducing the intake negative pressure as compared with a normal engine that does not perform stratified combustion, In the operation region A, the subsequent cylinders 2B and 2C have the air-fuel ratio substantially equal to the theoretical air-fuel ratio, and the compression self-ignition is performed in a uniform air-fuel mixture distribution state. The pumped loss is further reduced as compared with the preceding cylinders 2A and 2D because the gas thus supplied is sent. These effects greatly improve fuel efficiency.
[0073]
Moreover, since the gas discharged from the succeeding cylinders 2B and 2C into the exhaust passage 20 has a stoichiometric air-fuel ratio, it is not necessary to provide a lean NOx catalyst as in a conventional lean burn engine, and the three-way catalyst 24 is sufficient to exhaust the gas. Purification performance is ensured.
[0074]
Since there is no need to provide a lean NOx catalyst, there is no need to temporarily enrich the air-fuel ratio for NOx release and reduction when the NOx storage amount of the lean NOx catalyst is increased, thereby reducing fuel consumption improvement. can avoid. Furthermore, the problem of sulfur poisoning of the lean NOx catalyst does not occur.
[0075]
In addition, in the preceding cylinders 2A and 2D, the lean air-fuel ratio is set to approximately twice or close to the theoretical air-fuel ratio, so that the NOx generation amount can be suppressed to be relatively small. On the other hand, in the succeeding cylinders 2B and 2C, the burned gas is introduced from the preceding cylinders 2A and 2D so that a large amount of EGR is performed, and in the operation region A, compression auto-ignition occurs. Since the reaction between oxygen and nitrogen is avoided as much as possible when rapid combustion is performed, the generation of NOx is sufficiently suppressed. This is also advantageous for improving emissions.
[0076]
Further, in the operation region A1, which is the low and medium load region in the operation region A, in order to perform compression self-ignition reliably and smoothly, the valve timings of the intake valves 31 and the second exhaust valves 32b of the preceding cylinders 2A and 2D are conventionally set. The general valve timing of the engine is changed by the cam switching mechanism 51 as follows.
[0077]
FIG. 6 is a diagram showing the intake / exhaust stroke, fuel injection timing, ignition timing, etc. of each cylinder in the operation region A1, and FIG. 8 (a) shows the intake / exhaust stroke portion in detail.
[0078]
In FIG. 8A, the horizontal axis represents the crank angle, T represents the top dead center (TDC), and B represents the bottom dead center (BTC). The belt-shaped portions indicate the valve opening periods of the respective valves, and the white arrows from the upper stage to the lower stage indicate that the exhaust strokes of the preceding cylinders 2A and 2D and the succeeding cylinders 2B and 2C are substantially overlapped. , 2D shows a state where the burned gas is guided to the succeeding cylinders 2B, 2C. Further, the upper stage shows a valve opening period 80 in which the second exhaust valves 32b of the preceding cylinders 2A and 2D are opened, and a valve opening period 81 in which the intake valve 31 is opened, while the lower stage shows the subsequent cylinders 2B and 2C. The valve opening period 82 of the exhaust valve 32 and the valve opening period 83 during which the second intake valve 31b opens are shown.
[0079]
In the preceding cylinders 2A and 2D in the upper stage, the second cam 53 is selected by the cam switching mechanism 51, and the second exhaust valve 32b is more predetermined than the exhaust stroke top dead center 85 as clearly shown in FIG. The valve is closed before the period, and the burned gas remains in the combustion chamber 4. Then, fresh air is introduced into the preceding cylinders 2A and 2D by the supercharger 27, burned by forced ignition as described above, and then introduced into the succeeding cylinders 2B and 2C via the inter-cylinder passage 22. In the present embodiment, the valve timing of the second exhaust valve 32b is merely advanced from the general valve timing of the conventional engine, and the valve opening period or the like is the same as that of the conventional engine. It is set similarly to a general valve opening period.
[0080]
Thus, since fresh air is taken in and burned with the burned gas remaining in the preceding cylinders 2A and 2D, knocking in the preceding cylinders 2A and 2D is effectively prevented by increasing the burned gas component. Can do. In addition, since the preceding cylinders 2A and 2D take in fresh air into the relatively high temperature remaining burned gas and burn it, the burned gas becomes higher in temperature, and this high temperature burned gas is transferred to the succeeding cylinders 2B and 2C. Since it is introduced, the in-cylinder temperature in the succeeding cylinders 2B and 2C can be raised, and the compression self-ignition in the succeeding cylinders 2B and 2C is reliably and smoothly performed. Also, in the succeeding cylinders 2B and 2C, the burned gas of the preceding cylinders 2A and 2D is introduced, whereby the burned gas component corresponding to EGR is increased, thereby enhancing the knocking suppressing action of the succeeding cylinder.
[0081]
Here, in the preceding cylinders 2A and 2D, a part of the burned gas remains, and the intake valve 31 is opened in the exhaust stroke in order to overlap the second exhaust valve 32b. Although there is a concern about a decrease in output due to insufficient intake of air, a supercharger 27 is provided to assist the intake of the preceding cylinders 2A and 2D, so that the supercharger 27 provides fresh air to the preceding cylinders 2A and 2D. Can be forcibly sent in, and output reduction due to a reduction in intake air amount can be prevented.
[0082]
Then, as described above, as the second exhaust valves 32b in the preceding cylinders 2A and 2D are closed early, as shown in FIGS. 8A and 6, the intake valves 31 of the preceding cylinders 2A and 2D. Also, the second cam 53 is selected by the cam switching mechanism 51, the valve timing is changed, and the second cam 53 is overlapped with the second exhaust valve 32b. That is, even if the piston 3 is in the exhaust stroke, if the second exhaust valve 32b is closed early, the pumping loss increases and the fuel efficiency improvement effect is reduced. Since the exhaust valve 32b is overlapped, an increase in pumping loss can be avoided.
[0083]
Further, in the present embodiment, the valve timing is changed by the cam switching mechanism 51 as described above, and at the same time, the closing timing of the intake valve 31 is made to coincide with the intake stroke bottom dead center 87. Thus, by matching the valve closing timing of the intake valve 31 with the intake stroke bottom dead center 87, the effective compression ratio is made to coincide with the geometric compression ratio, and the increase in the in-cylinder temperature due to the compression is made larger. Thereby, higher-temperature burned gas can be introduced into the succeeding cylinders 2B and 2C. Therefore, the self-ignitability in the succeeding cylinders 2B and 2C can be reliably ensured even in a low load region where the in-cylinder temperature is unlikely to rise, and combustion by compression self-ignition in the succeeding cylinders 2B and 2C can be performed. The possible operating range can be expanded.
[0084]
As shown in FIGS. 8A and 6, the second intake valve 31b and the exhaust valve 32 in the succeeding cylinders 2B and 2C are selected by the cam switching mechanism 61, and the conventional engine is generally used. The valve is opened and closed at a typical valve timing.
[0085]
On the other hand, in the high load region A2 of the operation region A that is set to the special operation mode, the combustion temperature in the preceding cylinders 2A and 2D rises excessively and knocking easily occurs. And the valve timing of the 2nd exhaust valve 32b is changed into the general valve timing of the conventional engine.
[0086]
That is, FIG. 7 is a diagram showing the intake / exhaust stroke, fuel injection timing, ignition timing, etc. of each cylinder in the operation region A2, and the valve timing of each valve in FIG. 7 is a general valve timing of a conventional engine. Like the above, it is set as follows. FIG. 8B shows the intake / exhaust stroke portion of FIG. 7 in detail.
[0087]
In FIG. 8B, the horizontal axis represents the crank angle, T represents the top dead center (TDC), and B represents the bottom dead center (BTC). The belt-shaped portions indicate the valve opening periods of the respective valves, and the white arrows from the upper stage to the lower stage indicate that the exhaust strokes of the preceding cylinders 2A and 2D and the succeeding cylinders 2B and 2C are substantially overlapped. , 2D shows a state where the burned gas is guided to the succeeding cylinders 2B, 2C. The upper stage shows a valve opening period 100 in which the second exhaust valves 32b of the preceding cylinders 2A and 2D are opened and a valve opening period 101 in which the intake valve 31 is opened, while the lower stage shows the subsequent cylinders 2B and 2C. The valve opening period 102 of the exhaust valve 32 and the valve opening period 103 in which the second intake valve 31b opens are shown.
[0088]
In the high load region A2 of the operation region A that is set to the special operation mode, as described above, when the fresh burned gas is taken in and burned, the preceding cylinders 2A and 2D are likely to be knocked. As shown in FIGS. 7 and 8B, the intake valve 31 and the second exhaust valve 32b in the preceding cylinders 2A and 2D are selected by the cam mechanism 51 as the third cam 54, and the conventional engine is generally used. Open and close at valve timing.
[0089]
That is, the second exhaust valve 32b in the preceding cylinders 2A and 2D is set to be opened before a predetermined period before the exhaust stroke top dead center, and is closed after a predetermined period has passed after the exhaust stroke top dead center. And is opened for a predetermined period. For these predetermined periods, in the present embodiment, a general set value of a conventional engine is adopted as described above. As this set value, for example, the crank angle is opened about 10 ° before the exhaust stroke top dead center. After the exhaust stroke bottom dead center, the valve is closed at a crank angle of about 55 ° and opened at a crank angle of about 245 °.
[0090]
As shown in FIGS. 8B and 7, the second intake valve 31b and the exhaust valve 32 in the succeeding cylinders 2B and 2C are selected by the cam switching mechanism 61 so that the second intake valve 31b and the exhaust valve 32 in the operation region A1. As in the case, the valve is opened and closed at a general valve timing of a conventional engine.
[0091]
Thus, by switching the cams 52 to 54 according to the engine operating state by the cam switching mechanism 51, it is possible to set the optimum valve timing, and to ensure the output performance according to the engine operating state. it can. Further, for example, knocking in the succeeding cylinders 2B and 2C, which may be generated when the burned gas introduced into the succeeding cylinders 2B and 2C becomes excessively high in temperature, can be effectively prevented. That is, in the high load range, the temperature in the cylinders of both the preceding and succeeding cylinders 2A to 2D becomes high, and in this state, the preceding cylinders 2A and 2D take in fresh air into the burned gas and burn it. When the high-temperature burned gas from the preceding cylinders 2A and 2D is introduced into the succeeding cylinders 2B and 2C, the in-cylinder temperatures of the preceding and succeeding cylinders 2A to 2D become excessively high and abnormal combustion such as knocking occurs. May occur. Therefore, if comprised as mentioned above, while suppressing the temperature rise by the remaining burned gas in the preceding cylinders 2A and 2D, the burned gas temperature of the preceding cylinders 2A and 2D introduced into the succeeding cylinders 2B and 2C is lowered. Further, knocking in the preceding and succeeding cylinders 2A to 2D can be effectively prevented, and burned gas does not remain in the preceding cylinders 2A and 2D, so that the supercharging by the supercharger 27 is also helped and fresh air is sufficiently taken in. Therefore, the output performance of the engine can be fully exhibited.
[0092]
On the other hand, in the operation region B on the higher load side than the operation region A in the special operation mode, the normal operation mode is set, and the intake valve 31 is selected by the third cam 54 selected by the cam switching mechanism 51 as described above. Is activated, the second cam 63 is selected by the cam switching mechanism 61, the first exhaust valve 32a and the first intake valve 31a are activated, and the first cam 52 is selected by the cam switching mechanism 51. When the two exhaust valves 32b and the second intake valve 31b are stopped, the actual fresh air and gas flow paths are as shown in FIG. 10, and the intake ports 31, 31a of the cylinders 2A to 2D and The exhaust ports 12a and 12 are independent, and fresh air is introduced from the intake passage 15 to the intake ports 31 and 31a of the respective cylinders 2A to 2D by the supercharger 27, and the exhaust ports of the respective cylinders 2A to 2D. Burned gas is discharged into the exhaust passage 20 from 1,31A. In this case, the output performance is ensured by controlling the intake air amount and the fuel injection amount so that the stoichiometric air-fuel ratio or richer. Further, in this operation region B, the external EGR control valve 26 b is opened, and a part of the exhaust gas which is the burned gas of each cylinder flowing through the exhaust passage 20 is provided in this passage 26 via the external EGR passage 26. The EGR cooler 26b is introduced into the intake air passage 15 while being cooled. Thereby, it is possible to effectively prevent knocking by suppressing an increase in the in-cylinder temperature, and to secure the output performance of the engine corresponding to the load. The valve timing of each valve is set in the same manner as in the operation region A1.
[0093]
In addition, the specific structure of the apparatus of this invention is not limited to the said embodiment, A various change is possible. For example, the following changes are possible.
[0094]
(1) In addition to the control according to the operation state in the operation region A in the special operation mode as described above, the air-fuel ratio of the preceding cylinder may be changed according to the engine temperature state. For example, when the engine temperature is low even after the engine is warmed up (when the temperature of the engine cooling water is equal to or lower than a predetermined temperature), the preceding cylinder is emptied in the entire operation area A in the special operation mode. It is preferable to make the fuel ratio smaller than twice the theoretical air-fuel ratio. In this way, even when the engine temperature is relatively low, the temperature of the gas introduced from the preceding cylinder to the succeeding cylinder can be increased to ensure a state where compression self-ignition is possible.
[0095]
(2) In each of the above embodiments, the subsequent cylinder is burned by compression self-ignition in the operation region A1 on the low load side in the operation region A in the special operation mode. In part, for example, in an extremely low speed and low load region in which the temperature and pressure in the combustion chamber are difficult to reach a state in which compression self-ignition is possible, the ignition is performed by the spark plug 7 at a predetermined ignition timing for the subsequent cylinder, and forced ignition You may make it burn. Alternatively, when the engine temperature is low, the subsequent cylinder may be burned by forced ignition.
[0096]
(3) The apparatus of the present invention can be applied to multi-cylinder engines other than four-cylinder engines. For example, in the case of six cylinders, the exhaust stroke of one cylinder and the intake stroke of another cylinder do not completely overlap. In such a case, the exhaust stroke of one cylinder precedes the intake stroke of the other cylinder. In addition, two cylinders in which both strokes partially overlap may be used as a pair of preceding and succeeding cylinders.
[0097]
Next, another embodiment of the present invention will be described with reference to FIGS.
[0098]
In this other embodiment, the cylinders from the first cylinder 2A to the fourth cylinder 2D are independent from each other regardless of the operating state, and the valve timing variable mechanism for changing the valve timing in each cylinder. As a variable valve timing system (VTC) mechanism is adopted, the embodiment is greatly different from the above embodiment.
[0099]
FIG. 11 shows a schematic configuration of an engine according to another embodiment of the present invention, and FIG. 12 shows a configuration of a drive and control system in the device of this other embodiment.
[0100]
In addition, about the thing which is the same structure as the said embodiment in other embodiment, the same code | symbol is attached | subjected in a figure and the description is abbreviate | omitted.
[0101]
An intake port 111 and an exhaust port 112 are opened to the combustion chamber 4 of each cylinder 2A to 2D, and an intake passage 15 and an exhaust passage 20 are connected to these ports, and each port has an intake valve 131 and an exhaust valve. 132 is opened and closed.
[0102]
The intake / exhaust port of each cylinder and the intake passage, exhaust passage, and inter-cylinder gas passage connected to the cylinder are specifically configured as follows.
[0103]
As shown in FIG. 11, each of the cylinders 2 </ b> A to 2 </ b> D has an intake port 111 for introducing fresh air and an exhaust port 112 for sending exhaust gas to the exhaust passage 20. Two cylinders are arranged in parallel on each side. Further, the downstream end of the cylinder-specific branch intake passage 16 in the intake passage 15 is connected to the intake port 111 of each of the cylinders 2A to 2D.
[0104]
On the other hand, the intake / exhaust valves for opening and closing the intake / exhaust ports of each cylinder and the valve operating mechanisms for these valves are as follows.
[0105]
Each of these valves is operated at a predetermined timing by a valve mechanism according to the intake / exhaust stroke of each corresponding cylinder, that is, according to the vertical movement of each piston (movement between top dead center and bottom dead center). Although the opening and closing timing is not necessarily coincident with the top dead center or bottom dead center of the piston 3, the crank angle is advanced or retarded by several degrees to several tens degrees as necessary. It may be set at a different time. In particular, in this other embodiment, the valve timing can be changed according to the operating state of the engine by the variable valve timing mechanism provided in the valve operating mechanism.
[0106]
Specifically, the variable valve timing mechanism of this other embodiment is a variable valve timing mechanism that causes the valve timing to be advanced or retarded as a whole by relatively changing the rotational phase angles of the camshaft and the crankshaft. (VTC) is adopted. The cam phase variable mechanism 140 as the variable valve timing mechanism has a structure in which, for example, a helical gear built in one end of the camshaft 141 is hydraulically moved in the axial direction of the shaft to rotate the camshaft 141 relative to the pulley. This is a conventionally known mechanism.
[0107]
The cam phase varying mechanism 140 advances both the intake / exhaust valves 131 and 132 in the operation region corresponding to the operation region A1 (see FIG. 5) on the low load side, as shown in FIG. The exhaust valve 132 is set to be closed a predetermined period before the exhaust stroke top dead center, while the intake valve 131 is set to be closed to coincide with the intake stroke bottom dead center. Yes. Since the cam phase varying mechanism 140 changes only the cam phase, the intake valve 131 overlapped with the exhaust valve 132 is similarly overlapped after the change of the valve timing, The valve closing period or the like does not change before and after the valve timing is changed.
[0108]
On the other hand, in the operation region on the higher load side or higher rotation side than the operation region corresponding to the operation region A1, the intake / exhaust valves 131 and 132 are both retarded by the cam phase variable mechanism 140, as shown in FIG. Similar to the valve timing, the valve timing is changed to a general valve timing of a conventional engine. That is, in the operation region corresponding to the operation region A2 and the operation region B in the above-described embodiment, the valve opening period of the intake valve 131 is caused to straddle the intake stroke upper and lower dead centers corresponding to the intake stroke by the cam phase variable mechanism 140. On the other hand, the opening period of the exhaust valve 132 is set so as to straddle the upper and lower dead centers of the exhaust process corresponding to the exhaust stroke.
[0109]
In this embodiment, a variable valve timing system (VTC) is adopted as the variable valve timing mechanism. However, a known variable valve timing, such as a variable valve timing / lift mechanism (VVL) that changes the cam according to the operating state. It is the same as that of the said embodiment that a mechanism may be employ | adopted.
[0110]
FIG. 12 shows the configuration of the drive and control system. In this figure, an ECU (control unit) 140 for engine control comprising a microcomputer or the like is greatly different from the ECU 40 of the above embodiment in that a cam phase control means 142 is provided instead of the cam switching control means 42.
[0111]
The operation state discriminating means 141 includes an operation region corresponding to the operation region A1 shown in FIG. 5 in the above embodiment, and an operation region corresponding to the operation region A2 and the operation region B on the higher load side or higher rotation side than the operation region. The engine operating state (engine speed and engine load) checked by signals from the rotational speed sensor 47 and the accelerator opening sensor 48, etc., is divided into the low load side and the high load side. The operating region is determined, and the determination result is output to the combustion control means 144, the cam phase control means 142, and the like.
[0112]
The cam phase control unit 142 controls the cam phase variable mechanism 140 based on the result of the operation state determination unit 141. The cam phase control means 142 controls the cam phase variable mechanism 140 to advance the cam phase in a relatively low load region, and sets the opening / closing timings of the intake / exhaust valves 131 and 132 driven by the rotation of the camshaft 141. It is set to advance the whole. On the other hand, when the cam phase is advanced by the cam phase variable mechanism 140 and the cam phase enters the relatively high load region, the cam phase is returned, that is, the cam phase is delayed by the advanced amount. The cam phase variable mechanism 140 is controlled to the side, and the opening / closing timings of the intake / exhaust valves 131 and 132 driven by the rotation of the camshaft 141 are set to be retarded as a whole.
[0113]
The intake air amount control means 143 controls the opening degree (throttle opening degree) of the throttle valve 17 by controlling the actuator 18, obtains the target intake air amount according to the operating state, and obtains the target intake air. The throttle opening is controlled according to the amount. In this case, in the operation region on the relatively low load side, the combustion is performed while the ratio of the fresh air taken into each cylinder and the supplied fuel is set to the lean air-fuel ratio, so that the supercharger 27 is supercharged. Thus, the throttle opening is adjusted so that the air supplied to each cylinder exceeds the amount of air required for fuel combustion according to the required torque by a predetermined amount.
[0114]
The combustion control means 144 includes a fuel injection control means 145 and an ignition control means 146. The fuel injection control means 145 injects fuel from the fuel injection valves 9 provided corresponding to the cylinders 2A to 2D. The amount and the injection timing are controlled in accordance with 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. The control of the combustion state (control of fuel injection and control of ignition) is changed particularly when the operation state is in the operation region of the low load side and the high load side. That is, when the operating state is in the low load side operating region, the fuel injection amount is controlled so that the air-fuel ratio is set to a lean air-fuel ratio larger than the stoichiometric air-fuel ratio for each cylinder, and the fuel is injected during the intake stroke. The injection timing is set so as to be injected, and the forced ignition is stopped in order to perform the compression self-ignition. On the other hand, in the relatively high load operation region, the fuel injection amount is controlled so that the air-fuel ratio is the stoichiometric air-fuel ratio for each cylinder, and as in the case of FIG. The injection timing is set so that fuel is injected in the compression stroke, and the ignition timing is set so that forced ignition is performed in the vicinity of the compression top dead center.
[0115]
Next, the operation of the apparatus according to another embodiment will be described.
[0116]
In the apparatus of this other embodiment, the flow path of fresh air and gas is as shown in FIG. 11, and fresh air supplied from the branch intake passage 16 is burned in each cylinder and discharged to the branch exhaust passage 21. Independent cylinder state.
[0117]
In each cylinder, fresh air is introduced from the intake passage 15 during the intake stroke. ₂ While the fuel injection amount is feedback controlled so that the air-fuel ratio detected by the sensor 23 becomes a lean air-fuel ratio larger than the stoichiometric air-fuel ratio, after the fuel is injected in the intake stroke in the low load side operation region, the compression stroke is performed. In the vicinity of the top dead center, compression self-ignition is performed by an increase in pressure and temperature in the combustion chamber (see FIG. 13). On the other hand, in the high-load operation region, after the fuel is injected in the compression stroke, Near the top dead center, combustion at the stoichiometric air-fuel ratio is performed by forced ignition.
[0118]
As described above, in each of the cylinders 2A to 2D, in the operation region on the low load side, the heat efficiency is improved by lean combustion, and the pumping loss is reduced by reducing the intake non-pressure as compared with a normal engine that does not perform lean. Is done. These effects greatly improve fuel efficiency.
[0119]
Further, in each of the cylinders 2A to 2D in the operation region on the low load side, compression self-ignition is performed in a uniform mixture distribution state, so that combustion is performed all over the combustion chamber 4 so that slow combustion that does not contribute to work is avoided. Therefore, a high fuel efficiency improvement effect can be obtained by improving thermal efficiency.
[0120]
Then, in order to perform compression self-ignition reliably and smoothly in the low load side operation region, the exhaust valve 132 of each cylinder 2A to 2D is advanced by the cam phase variable mechanism 140 as a whole, and the exhaust stroke top dead center. Rather, the burned gas remains in the combustion chamber 4. Since this burned gas is a relatively high temperature, the in-cylinder temperature also becomes high, and the compression self-ignition is surely and smoothly performed in combination with the fresh air introduced into the cylinders 2A to 2D by the supercharger 27. Is called. In addition, since the EGR is performed with the remaining burned gas, the amount of NOx generated is sufficiently suppressed to contribute to the purification of exhaust gas, and the EGR in the gas is reduced. The knocking suppression effect is enhanced by increasing the corresponding burned gas component.
[0121]
On the other hand, as the exhaust valve 132 is closed early, the intake valve 131 is also advanced by the cam phase varying mechanism 140 and overlapped with the exhaust valve 132. That is, when the exhaust valve 132 is closed early even though the piston 3 is in the exhaust stroke, the pumping loss increases and the fuel efficiency improvement effect is reduced. In this way, the intake valve 131 is exhausted. By overlapping the valve 132, an increase in pumping loss is avoided.
[0122]
Further, the closing timing of the intake valve 131 is made to substantially coincide with the bottom dead center of the intake stroke, the effective compression ratio is made to coincide with the geometric compression ratio, and the increase in the in-cylinder temperature due to the compression is larger. Thus, the in-cylinder temperature can be maintained at a higher temperature. Therefore, the self-ignitability in each of the cylinders 2A to 2D can be reliably ensured even in a low load range where the in-cylinder temperature is unlikely to rise, and combustion by compression self-ignition in each of the cylinders 2A to 2D can be performed. The possible operating range can be expanded.
[0123]
Here, a part of the burned gas remains in the cylinders 2A to 2D, and the intake valve 131 is opened in the exhaust stroke in order to overlap the exhaust valve 132. Although there is a concern about a decrease in output due to insufficient intake, a supercharger is provided to assist the intake of each cylinder 2A to 2D. Therefore, this supercharger forces fresh air to each cylinder 2A to 2D. Therefore, the output can be prevented from decreasing due to a decrease in the intake air amount.
[0124]
On the other hand, when the operation region shifts from the low load side operation region to the high load side operation region, the intake / exhaust valves 131 and 132 are retarded by the cam phase variable mechanism 140, and the valve timing is the original valve timing. Changed to timing. Further, in each of the cylinders 2A to 2D, the combustion control means 144 shifts from combustion by compression self-ignition to combustion by forced ignition to prevent knocking due to an excessive rise in the in-cylinder temperature.
[0125]
In the engine according to the other embodiment, the fuel injection valve 9 is provided at a side portion of the combustion chamber 4 and directly injects fuel into the combustion chamber 4. It may be provided so that fuel is injected into the fresh air in the intake port 111 and the air-fuel mixture is sent into the combustion chamber 4.
[0126]
【The invention's effect】
As described above, according to the control device of the first aspect of the present invention, when the special operation mode is set, combustion is performed at the lean air-fuel ratio in the preceding cylinder of the pair of cylinders in which the exhaust stroke and the intake stroke overlap, 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 by compression self-ignition. Therefore, in the preceding cylinder, thermal efficiency is improved by lean combustion and pumping loss is reduced. In the succeeding cylinder, the fuel efficiency can be improved by improving the combustion efficiency by compression self-ignition and reducing the pumping loss. In addition, since the air-fuel ratio at the time of combustion in the subsequent cylinder is substantially the stoichiometric air-fuel ratio, the exhaust gas in the exhaust passage can be sufficiently purified with only the three-way catalyst, and the lean NOx catalyst is It becomes unnecessary.
[0127]
In addition, since fresh air is taken in and burned with the burned gas remaining in the preceding cylinder, the knocking suppression effect in the preceding cylinder is enhanced, and the burned gas in the preceding cylinder becomes even higher, and this high temperature burned gas is increased. The combustion gas is introduced into the succeeding cylinder so that the compression self-ignition in the succeeding cylinder can be performed reliably and smoothly. Also in the subsequent cylinder, the knocking suppression action is enhanced by increasing the burned gas component corresponding to EGR. On the other hand, as the exhaust valve of the preceding cylinder is closed early, the intake valve is overlapped with the exhaust valve, so an increase in pumping loss can be avoided.
[0128]
Further, although the burned gas remains in the preceding cylinder, since the fresh air is sent to the preceding cylinder by the supercharger, it is possible to prevent a decrease in output due to a decrease in the intake air amount.
[0129]
On the other hand, according to the apparatus of the invention described in claim 6, since the entire combustion chamber is combusted by compression self-ignition, slow combustion that does not contribute to work is avoided, and thermal efficiency is improved and a high fuel efficiency improvement effect is obtained. Moreover, since it will be in the state equivalent to EGR being performed with residual burnt gas, it will contribute to purification | cleaning of waste gas by fully suppressing the generation amount of NOx. Furthermore, since fresh air is taken in and burned in the relatively high-temperature burned gas remaining in the cylinder, the in-cylinder temperature can be raised, and thereby compression self-ignition in the cylinder can be performed reliably and smoothly. . Moreover, the knocking suppression action is also enhanced in the subsequent cylinders due to an increase in the burned gas component corresponding to EGR. On the other hand, as the exhaust valve is closed early, the intake valve is overlapped with the exhaust valve, so an increase in pumping loss can be avoided. Further, the reduction of output due to insufficient intake of fresh air can be reliably prevented by feeding fresh air by the supercharger.
[Brief description of the drawings]
FIG. 1 is a schematic plan view of an entire engine including a control device according to an embodiment of the present invention.
FIG. 2 is a schematic cross-sectional view of an engine body and the like.
FIG. 3 is a block diagram of a control system.
FIG. 4 is a partial perspective view showing a cam switching mechanism.
FIG. 5 is an explanatory diagram showing an example of operation region setting for performing control according to an operation state.
FIG. 6 is a diagram illustrating an exhaust stroke, an intake stroke, a fuel injection timing, an ignition timing, and the like of each cylinder on a low load side during a special operation mode.
FIG. 7 is a diagram showing an exhaust stroke, an intake stroke, a fuel injection timing, an ignition timing, and the like of each cylinder on a high load side during a special operation mode.
8A is a detailed view of FIG. 6 and FIG. 8B is a detailed view of FIG. 7;
FIG. 9 is an explanatory diagram showing a substantial fresh air and gas flow path in a special operation mode.
FIG. 10 is an explanatory diagram showing substantial fresh air and gas flow paths in the normal operation mode.
FIG. 11 is an overall schematic plan view including a control device according to another embodiment.
FIG. 12 is a block diagram of a control system according to another embodiment.
FIG. 13 is a diagram showing an exhaust stroke, an intake stroke, a fuel injection timing, an ignition timing, and the like of each cylinder on the low load side according to another embodiment.
[Explanation of symbols]
1 Engine body
2A to 2D cylinder
15 Intake passage
20 Exhaust passage
22 Gas passage between cylinders
26 External EGR passage
27 Turbocharger
31 Intake valve
32 Exhaust valve
40 ECU
41 Operating state discriminating means
42 Cam switching control means
43 Intake air amount control means
44 Combustion control means
51 Cam switching mechanism (variable valve timing mechanism)
41 Operating state discriminating means

Claims (5)

In a multi-cylinder spark-ignition four-cycle engine in which the combustion cycle of each cylinder is performed with a predetermined phase difference, the engine intake / exhaust and combustion state control modes are specially operated in the engine partial load range. In this special operation mode, the burned gas discharged from the preceding cylinder in the exhaust stroke between a pair of cylinders in which the exhaust stroke and the intake stroke overlap is directly passed to the subsequent cylinder in the intake stroke via the inter-cylinder gas passage. The two cylinders are connected so that the gas discharged from the succeeding cylinder is introduced into the exhaust passage, while the preceding cylinder performs combustion at a lean air / fuel ratio larger than the stoichiometric air / fuel ratio. A control device for a spark ignition engine in which fuel is supplied to burned gas having a lean air-fuel ratio introduced from a preceding cylinder to cause combustion. ,
In at least a part of the operation region set in the special operation mode, in the preceding cylinder, the exhaust valve is closed before the exhaust stroke top dead center so that the burned gas is kept in the cylinder. While controlling the valve opening period of the intake valve to overlap the valve opening period of the exhaust valve, and controlling the subsequent cylinder to perform combustion by compression self-ignition,
A control device for a spark ignition engine equipped with a supercharger, wherein a supercharger for assisting intake in the preceding cylinder is provided.   2. The control device for a spark ignition engine with a supercharger according to claim 1, wherein the closing timing of the intake valve in the preceding cylinder is set to substantially coincide with the bottom dead center of the intake stroke. Control device for a spark ignition engine with a supercharger.   The control apparatus for a spark ignition engine with a supercharger according to claim 1 or 2, further comprising a valve timing variable mechanism for changing a valve timing of an intake / exhaust valve in the preceding cylinder, wherein the special operation mode is set. In a higher load range than the partial operation region in the operation region, the exhaust valve of the preceding cylinder is set to be closed after the exhaust stroke top dead center has elapsed for a predetermined period by the variable valve timing mechanism. In addition, the control device for a spark ignition engine with a supercharger is set so that the intake valve of the preceding cylinder is closed after a predetermined period of time has passed after the intake stroke bottom dead center.   4. The control apparatus for a spark ignition engine with a supercharger according to claim 3, wherein the intake / exhaust valve in the preceding cylinder is set to its intake / exhaust stroke in a higher load region than the operation region in which the special operation mode is set. It is set to switch to the normal operation mode in which each cylinder is independently opened and closed by forced ignition, and in this normal operation mode, part of the exhaust gas in each cylinder is provided with an EGR cooler. A control device for a spark ignition engine with a supercharger, wherein the control device is introduced into an intake passage through an external EGR passage.   The supercharger-equipped spark ignition engine control device according to any one of claims 1 to 4, wherein the supercharger is a turbocharger. Control device for ignition engine.

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