JP5310951B2 - Control device for internal combustion engine - Google Patents

Abstract

The purpose of the present invention is to prevent variations in combustion between cycles and provide stabilized combustion by injecting fuel mainly through an intake port through which a reduced flow of air is allowed. An engine (10) is provided with first and second intake ports (20A, 20B), first and second exhaust ports (22A, 22B), first and second fuel injection valves (24A, 24B), and an SCV (36). In the intake stroke, while the first intake port (20A) is blocked with the SCV (36), the second intake port (20B) allows air to flow into the cylinder and the first fuel injection valve (24A) injects fuel into the intake port (20A). Since this allows the injected fuel not to be significantly affected by the air flow in the first intake port (20A), the injected fuel can be fed into the cylinder while maintaining a natural spray shape. It is thus possible to prevent variations in combustion between cycles which would otherwise result from disturbance of the spray shape.

Description

  The present invention relates to a control device for an internal combustion engine, and more particularly to a control device for an internal combustion engine having a plurality of intake ports in one cylinder.

As a conventional technique, for example, as disclosed in Patent Document 1 (Japanese Patent Laid-Open No. 2003-314339), each cylinder is provided with two intake ports, and a fuel injection valve and an intake control valve are provided in each intake port. A control device for an arranged internal combustion engine is known. In the prior art, intake control valves having different shapes (different shut-in areas of the intake ports) are arranged in the two intake ports in each cylinder. During operation of the internal combustion engine, fuel is injected into the intake port on the side where the intake control valve is open with one intake control valve closed and the other intake control valve opened. . As a result, in the prior art, the injection amount at each intake port is individually controlled according to the opening of the intake control valve to stabilize combustion.
The applicant has recognized the following documents including the above-mentioned documents as related to the present invention.

Japanese Unexamined Patent Publication No. 2003-314339 Japanese Unexamined Patent Publication No. 2006-144752 Japanese Unexamined Patent Publication No. 2009-062946 Japanese Unexamined Patent Publication No. 2008-291686 Japanese Patent Laid-Open No. 06-241054 Japanese Unexamined Patent Publication No. 2009-103108 Japanese Unexamined Patent Publication No. 2004-251143 Japanese Unexamined Patent Publication No. 2001-323828 Japanese Unexamined Patent Publication No. 06-323168

  By the way, in the above-described prior art, fuel is injected into the intake port on the side where the intake control valve is opened, and the injected fuel and air are synchronized and introduced into the cylinder. However, the air flow generated in the intake port changes in a complicated manner depending on the engine speed, the load, the opening degree of the intake control valve, and the like. Therefore, when fuel is injected into this air flow, the fuel spray shape and the state of mixing with air change irregularly, and the amount and position of fuel adhering to the wall surface of the intake port and the intake valve accordingly. Etc. fluctuate. For this reason, the conventional technology has a problem that even when the fuel injection amount and the injection timing are fixed, a phenomenon in which the combustion state of the air-fuel mixture fluctuates for each cycle (combustion variation between cycles) occurs and exhaust emission deteriorates. is there.

  The present invention has been made to solve the above-described problems, and an object of the present invention is to suppress combustion variation between cycles of an internal combustion engine having a fuel injection valve in an intake port, thereby stabilizing combustion. An object of the present invention is to provide a control device for an internal combustion engine that can be made to operate.

1st invention, 1st 2nd intake port which the upstream side is connected to an intake passage, and the downstream side opens in a pipe in a different position, respectively,
First and second intake valves provided in the first and second intake ports;
First and second fuel injection valves for injecting fuel into the first and second intake ports;
An intake adjustment valve that is provided in the first intake port upstream of the first fuel injection valve and changes a flow passage area of the first intake port;
At least in the intake stroke, the intake adjustment valve is driven to the closed side to reduce the air flow rate of the first intake port to be lower than the air flow rate of the second intake port, and the first fuel injection valve Injection state control means for performing fuel injection in a state where the fuel injection amount is set to be larger than the fuel injection amount of the second fuel injection valve.

  According to the second invention, the injection state control means closes the intake adjustment valve and shuts off the air flow in the first intake port, and the injection state control means during the opening period of the first intake valve. The fuel is injected by the first fuel injection valve.

  3rd invention is provided with the division | segmentation injection means which divides | segments and injects the injection fuel for 1 cycle injected by the said 1st fuel injection valve in multiple times.

A fourth invention is a variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves for each valve;
Valve closing timing control means for delaying the closing timing of the second intake valve from the closing timing of the first intake valve.

5th invention, the injection state change means which performs a fuel injection in the state which set more fuel injection quantity of the said 2nd fuel injection valve than the fuel injection quantity of the said 1st fuel injection valve,
Control switching means for operating the injection state changing means instead of the injection state control means when raising the temperature of the first intake valve.

  According to a sixth invention, the control switching means operates the injection state changing means when the air-fuel ratio of the internal combustion engine is close to the stoichiometric air-fuel ratio, and the air-fuel ratio is leaner than the stoichiometric air-fuel ratio. In this case, the injection state control means is operated.

A seventh aspect of the present invention is a lift amount variable mechanism capable of variably setting a lift amount of the second intake valve;
And a starting lift amount reducing means for reducing the lift amount of the second intake valve compared with the lift amount after starting when the internal combustion engine is started.

The eighth invention is a shut-off valve capable of closing the intake port except for a portion of the second intake port that allows air to flow into the outer periphery of the cylinder, at least in the latter half of the intake stroke. Equipped with a driven intake partial shut-off valve,
The injection state control means is configured to inject fuel into the first half of the intake stroke by the first fuel injection valve.

A ninth invention is a variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves for each valve;
Valve closing timing control means for delaying the closing timing of the second intake valve from the closing timing of the first intake valve.

A tenth aspect of the invention includes an ignition retarding means for retarding an ignition timing when starting an internal combustion engine,
The injection state control means is configured to inject fuel in the first half and the second half of the intake stroke by the first fuel injection valve when the internal combustion engine is started.

An eleventh aspect of the present invention is the first and second exhaust ports provided at positions facing the first and second intake ports,
First and second exhaust valves provided in the first and second exhaust ports;
A variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves and the first and second exhaust valves for each valve;
Of the first and second intake valves and the first and second exhaust valves, only the first and second intake valves are opened in the first half of the intake stroke, and only the first exhaust valve is opened. And a valve opening period control means that opens in the latter half of the intake stroke,
The injection state control means is configured to inject fuel into the first half of the intake stroke by the first fuel injection valve.

In a twelfth aspect, the first and second exhaust ports provided at positions facing the first and second intake ports,
First and second exhaust valves provided in the first and second exhaust ports;
A variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves and the first and second exhaust valves for each valve;
Valve opening period control means for opening the first and second intake valves over the first half and the second half of the intake stroke, and opening the first exhaust valve in the second half of the intake stroke;
The injection state control means is configured to inject fuel in the first half and the second half of the intake stroke by the first fuel injection valve.

  A thirteenth aspect of the invention is a shut-off valve capable of closing the intake port except for a portion corresponding to the inner and outer circumferences of the second intake port, and is driven to the closed side at least in the first half of the intake stroke. An intake partial shutoff valve is provided.

  A fourteenth aspect of the invention is a shut-off valve capable of closing the exhaust port except for a portion corresponding to the inner and outer circumferences of the first exhaust port, and is driven to the closing side in the latter half of the intake stroke. An exhaust partial shutoff valve is provided.

A fifteenth aspect of the invention is a variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves for each valve;
An opening timing control means for delaying the opening timing of the second intake valve from the opening timing of the first intake valve in the intake stroke;
The injection state control means is in a state where the first intake valve is opened and the second intake valve is opened after the first intake port is shut off by the intake adjustment valve. The fuel is injected by the first fuel injection valve.

The sixteenth invention comprises an exhaust valve provided at the exhaust port;
A variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves and the exhaust valves for each valve;
Closing the exhaust valve before exhaust top dead center, opening the first intake valve between exhaust valve closing and exhaust top dead center, and And an opening / closing timing control means for opening the intake valve of 2 after exhaust top dead center,
The injection state control means is configured such that the first intake port is shut off by the intake adjustment valve and the first intake valve is opened until the exhaust top dead center arrives. The fuel is injected by one fuel injection valve.

According to a seventeenth aspect, the injection state control means is
Injection means before and after top dead center for injecting fuel before and after exhaust top dead center by the first fuel injection valve;
If the engine speed is above the reference value, increase the fuel injection amount before exhaust top dead center over the fuel injection amount after exhaust top dead center, and if the engine speed is below the reference value, Injection amount control means for increasing the fuel injection amount after the dead center more than the fuel injection amount before the exhaust top dead center.

  The eighteenth aspect of the invention comprises self-ignition combustion control means for self-igniting the air-fuel mixture in the cylinder by compression.

In a nineteenth aspect of the invention, the first intake port and the first fuel injection valve are arranged on one side with respect to a reference straight line passing through the center of the combustion chamber in a plan view, and the second intake port and The second fuel injection valve is disposed on the other side with respect to the reference straight line,
At least the fuel injection direction of the first fuel injection valve is inclined toward the center of the combustion chamber.

A twentieth aspect of the invention includes first and second exhaust ports provided at positions facing the first and second intake ports in a direction parallel to the reference straight line.
First and second exhaust valves provided in the first and second exhaust ports,
The fuel injection angle, which is an angle at which the fuel injected from the first fuel injection valve expands in plan view, is such that the center of the second intake valve and the center of the first exhaust valve are inside the fuel injection angle. The angle is set to be within the range.

  In a twenty-first aspect, the center line of the fuel spray injected from the first fuel injection valve is disposed so as to pass through a position closer to the second fuel injection valve than the center of the combustion chamber in a plan view. It is configured to do.

  According to a twenty-second aspect, the injection state control means sets the fuel injection period of the first fuel injection valve to the first half of the intake stroke.

  A twenty-third aspect of the invention is a means that operates in place of the injection state control means when the internal combustion engine is operated in a high rotation and high load state, and the first and first intake valves are opened when the intake adjustment valve is opened. High-power control means for performing fuel injection by the second fuel injection valve is provided.

  According to a twenty-fourth aspect of the invention, the injection state control means drives the intake adjustment valve to the open side and increases the fuel injection amount of the second fuel injection valve as the load of the internal combustion engine increases, The difference between the fuel injection amounts by the first and second fuel injection valves is reduced.

In a twenty-fifth aspect of the invention, first and second intake ports that are connected to the intake passage on the upstream side and open into the cylinder at different positions on the downstream side,
First and second fuel injection valves for injecting fuel into the first and second intake ports;
When the air flow rate of the first intake port is smaller than the air flow rate of the second intake port in the intake stroke, the fuel injection amount of the first fuel injection valve is set to the fuel injection amount of the second fuel injection valve. And an injection state control means for executing fuel injection in a state set to be larger than the amount.

  According to the first aspect of the present invention, when the intake adjustment valve is driven to the closed side, it is noted that the amount of air passing through the first intake port by the intake adjustment valve is smaller than that of the second intake port. The amount of fuel injected into the first intake port is increased. Thus, the spray shape of the fuel injected into the intake port is prevented from being disturbed by the air flow, and the amount and position of fuel adhering to the surrounding structure can be stabilized. Therefore, variation in combustion between cycles due to disturbance of the spray shape can be suppressed, and for example, an operation region in which lean combustion can be performed can be expanded, and exhaust emission can be improved. Further, a swirl flow can be formed in the cylinder by mainly flowing air into the cylinder from the second intake port. This swirl flow improves the homogeneity of the air-fuel mixture, and the combustion speed can be improved by improving the turbulence of the in-cylinder airflow.

  According to the second aspect of the present invention, the air flow in the first intake port can be blocked by the intake adjustment valve, and a state where the air in the intake port is stably retained can be realized. Thus, the injected fuel can enter the cylinder from the opening of the intake port without being affected by the air flow.

  According to the third aspect of the invention, the injected fuel for one cycle can be divided into several times and injected in small amounts, so that the vaporization of the injected fuel can be promoted and the homogeneity of the air-fuel mixture can be improved. Therefore, the combustion speed of the air-fuel mixture can be further improved, the limit of the air-fuel ratio (lean limit) at which lean combustion is possible can be expanded, and the generation of HC due to unburned fuel can be suppressed.

  According to the fourth invention, the so-called Atkinson cycle can be realized by reducing the intake air amount by delaying the closing timing of the second intake valve. Thereby, thermal efficiency can be improved and fuel consumption can be improved. On the other hand, since the closing timing of the first intake valve can be advanced, it is possible to prevent the injected fuel from blowing back from the inside of the cylinder to the first intake port. As a result, the injected fuel that is blown back is not carried over to the next cycle, so the responsiveness during acceleration and deceleration operations can be improved, and combustion variations and torque fluctuations between cycles can be reduced. .

  According to the fifth aspect, the control switching means can operate the injection state changing means in place of the injection state control means. As a result, a state where a large amount of fuel is not injected in the vicinity of the first intake valve can be realized, and the first intake valve can be prevented from being cooled by the injected fuel. Therefore, the temperature of the first intake valve can be increased as necessary, and deposits can be suppressed from adhering to the first intake valve.

  According to the sixth aspect of the invention, when the air-fuel ratio is in the vicinity of stoichiometric, the combustion temperature tends to increase. Therefore, by operating the injection state changing means at the timing when the air-fuel ratio becomes stoichiometric, the first intake valve can be efficiently heated in a short time using this timing. As a result, it is possible to obtain a deposit suppressing effect during stoichiometric combustion while operating the injection state control means during lean combustion.

  According to the seventh aspect, at the time of start-up, the flow area of the second intake port can be reduced according to the valve lift amount, and the flow rate of air flowing into the cylinder from the intake port can be increased. As a result, the swirl flow in the cylinder can be strengthened in the direction opposite to the fuel injection direction, the momentum of the injected fuel can be reduced by this swirl flow, and the injected fuel can be diffused efficiently. As a result, vaporization of the injected fuel can be promoted even during a cold start, and the fuel can be prevented from adhering to the cylinder inner wall surface. Therefore, the startability of the internal combustion engine and the exhaust emission at the start can be improved.

  According to the eighth aspect, in the first half of the intake stroke, an air-fuel mixture is formed in the cylinder by the fuel injected from the first fuel injection valve and the air flowing in from the second intake port. Further, in the latter half of the intake stroke, only air flows from the second intake port into the cylinder outer periphery due to the action of the intake partial shutoff valve. As a result, an air-fuel mixture layer can be formed in the center of the cylinder and an air layer can be formed on the outer periphery of the cylinder, so that stratified combustion can be realized. Accordingly, combustion can be stabilized even with a small amount of fuel injection, and lean combustion can be easily performed. Moreover, pump loss and cooling loss can be suppressed, and fuel consumption can be improved.

  According to the ninth aspect, the effect of the Atkinson cycle can be obtained as in the fourth aspect. In addition, in the Atkinson cycle, there is a possibility that the air-fuel mixture in the cylinder blows back to the second intake port due to the late closing timing of the second intake valve. On the other hand, in the ninth invention, even if blowback to the second intake port occurs, the blown back gas becomes an air layer on the outer periphery in the cylinder. Therefore, it is possible to suppress the air-fuel mixture amount in the cylinder from fluctuating due to blow-back and reduce the combustion variation between cycles.

  According to the tenth aspect of the invention, since the air-fuel mixture layer is formed in the vicinity of the spark plug, the air-fuel mixture layer can improve the ignitability at the start. Further, the fuel injection in the latter half of the intake stroke can form an air-fuel mixture also on the inner and outer periphery of the cylinder, and can increase the combustion amount of the air-fuel mixture and raise the exhaust gas temperature. Furthermore, the exhaust gas temperature can be increased by retarding the ignition timing. Therefore, the startability of the internal combustion engine can be improved, and the catalyst can be quickly warmed up even during a cold start.

  According to the eleventh aspect, an air-fuel mixture can be formed in the cylinder in the first half of the intake stroke, and exhaust gas can be introduced into the cylinder from the first exhaust port in the second half of the intake stroke. At this time, since the first exhaust port is not opposed to the second intake port that introduces air into the cylinder in the first half of the intake stroke, the exhaust gas can be introduced into the outer periphery of the cylinder. As a result, it is possible to realize a stratified EGR in which an air-fuel mixture layer is formed in the center of the cylinder and an exhaust gas (EGR gas) layer is formed on the outer periphery of the cylinder. Therefore, fuel efficiency can be improved by stratified combustion. Further, EGR can be executed, the air-fuel ratio can be maintained at the stoichiometric air-fuel ratio, and NOx in the exhaust gas can be reduced by the three-way catalyst.

  According to the twelfth invention, the stratified EGR can be realized in substantially the same manner as the eleventh invention. In addition, since the second intake valve is opened in the second half of the intake stroke and fuel injection is also executed in the second half of the intake stroke, a gas layer in which exhaust gas and air-fuel mixture are mixed is formed on the outer periphery of the cylinder. can do. This gas layer is difficult to burn by flame propagation from the central part of the cylinder, but burns by self-ignition by the heat of exhaust gas contained in itself in the latter half of the combustion stroke. Therefore, the unburned fuel that tends to remain on the outer periphery of the cylinder can be reliably burned using the heat of the exhaust gas introduced into the cylinder. Thereby, HC in exhaust gas can be suppressed and exhaust emission can be improved.

  According to the thirteenth aspect, the intake partial shutoff valve can strengthen the swirl flow in the cylinder in the first half of the intake stroke, and can bring the position where the swirl flow flows closer to the cylinder inner wall surface. Thereby, in the second half of the intake stroke, exhaust gas can be efficiently guided to the inner and outer peripheries using a strong swirl flow, and turbulence of the stratified EGR can be reduced.

  According to the fourteenth aspect of the invention, the exhaust partial shut-off valve can force exhaust gas to flow into the cylinder outer periphery in the second half of the intake stroke, and prevent the exhaust gas from flowing into the cylinder center. be able to. Therefore, the stratified EGR can be realized stably.

  According to the fifteenth aspect, at the time of fuel injection, the piston performs the lowering operation with the first intake port blocked by the intake adjustment valve and the other ports closed by the valves. As a result, fuel can be injected into the cylinder in a negative pressure state, so that the boiling point of the fuel can be lowered and vaporization of the injected fuel can be promoted. And a favorable air-fuel mixture can be formed quickly, and the injected fuel can be prevented from adhering to the inner wall surface of the cylinder.

  According to the sixteenth aspect, during the period from when the first intake valve is opened to when the second intake valve is opened, the first intake port is blocked by the intake adjustment valve, Each port is closed by a valve. As a result, the inside of the cylinder is hermetically sealed, so that the exhaust gas remaining in the cylinder is once compressed before the exhaust top dead center and becomes high temperature. When the fuel is injected in this state, the injected fuel is reformed into a radical state or component that easily ignites in high-temperature exhaust gas. Therefore, the injected fuel can be easily burned by the fuel reforming effect, and the ignitability of the air-fuel mixture can be improved. Moreover, vaporization of the injected fuel can be promoted by the high-temperature exhaust gas, and the fuel can be prevented from adhering to the inner wall surface of the cylinder.

  According to the seventeenth aspect, at the time of high revolution, the fuel injection amount before exhaust top dead center can be increased, and a sufficient fuel reforming effect can be obtained even if the time from fuel injection to ignition is short. it can. Moreover, at the time of low rotation, the fuel injection amount after exhaust top dead center can be increased, and both the fuel reforming effect and the homogeneity of the air-fuel mixture can be achieved. Therefore, the fuel reforming level can be appropriately controlled according to the engine speed.

  According to the eighteenth aspect, fuel reforming can be performed by injecting fuel into the high-temperature exhaust gas once compressed in the cylinder. Thereby, at the time of execution of the self-ignition combustion control, the injected fuel can be stably self-ignited even in the low load operation region, and misfires can be prevented.

  According to the nineteenth aspect, the first fuel injection valve can inject fuel toward the center of the combustion chamber. As a result, most of the injected fuel does not collide with the swirl flow flowing along the outer periphery of the cylinder from the front, so that the swirl flow can be prevented from being weakened by the fuel spray. Therefore, since the swirl flow can be maintained until immediately before combustion, the homogenization of the air-fuel mixture can be promoted by the turbulent air flow, and the combustibility and fuel consumption can be improved. Further, the injected fuel can be injected in the diameter direction of the combustion chamber, that is, in the direction in which the distance to the cylinder inner wall surface becomes the longest, and the fuel can be prevented from adhering to the cylinder inner wall surface.

  According to the twentieth invention, most of the in-cylinder space can be accommodated within the range of the fuel injection angle by the first fuel injection valve. Thereby, the injected fuel can be dispersed in a wide range in the cylinder, fuel vaporization can be promoted, and a homogeneous air-fuel mixture can be formed throughout the cylinder.

  According to the twenty-first aspect, the fuel injection direction can be matched to the direction of the swirl flow flowing along the outer periphery in the cylinder. Thereby, a swirl flow can be strengthened by spraying of fuel, and a swirl flow can be maintained until immediately before combustion. Therefore, the homogenization of the air-fuel mixture can be promoted by the turbulent air flow, and the combustibility and fuel consumption can be improved.

  According to the twenty-second aspect of the present invention, the swirl flow can be accelerated by fuel spray at the first half of the intake stroke when the swirl flow generated in the cylinder is relatively weak. Therefore, the swirl flow in the cylinder can be strengthened from the first half of the intake stroke, and the effect of the twenty-first invention can be exhibited more remarkably. In addition, when fuel is injected into a strong swirl flow, the fuel droplets are likely to adhere to the inner wall surface of the cylinder due to centrifugal force before vaporizing the fuel. Therefore, by injecting the fuel when the swirl flow is weak, it is possible to prevent the injected fuel from adhering to the inner wall surface of the cylinder.

  According to the twenty-third aspect of the present invention, during the high rotation and high load operation, the fuel injection is executed by the first and second fuel injection valves with the intake adjustment valve opened, and the intake air amount and the fuel injection amount are sufficient. Can be secured. Therefore, the control can be appropriately switched according to the operating state.

  According to the twenty-fourth aspect, the injection state control means can drive the intake adjustment valve to the open side and increase the fuel injection amount of the second fuel injection valve as the load of the internal combustion engine increases. Thereby, even at the time of high load operation, the intake air amount and the fuel injection amount can be sufficiently secured, and the drivability can be improved.

  According to the twenty-fifth aspect, in the intake stroke, fuel injection can be executed when the air flow rate of the first intake port is small. As a result, the fuel injected from the first fuel injection valve is not greatly affected by the air flow in the first intake port, so that the fuel spray shape can be prevented from being disturbed by the air flow. it can. Therefore, variation in combustion between cycles can be suppressed and exhaust emission can be improved. In addition, by causing air to flow mainly into the cylinder from the second intake port, a swirl flow can be formed in the cylinder and the combustion speed can be improved.

It is a whole block diagram for demonstrating the system configuration | structure of Embodiment 1 of this invention. It is a block diagram which expands and shows the structure of 1 cylinder of an engine. It is explanatory drawing which shows the effect obtained by injection fuel stabilization control. In Embodiment 1 of this invention, it is a flowchart which shows the control performed by ECU. It is explanatory drawing which shows the division | segmentation injection of the fuel in Embodiment 2 of this invention. In Embodiment 3 of this invention, it is explanatory drawing which shows the opening / closing timing of the 1st, 2nd intake valve. In Embodiment 4 of this invention, it is a flowchart which shows the control performed by ECU. In Embodiment 5 of this invention, it is explanatory drawing which shows the effect of the control which reduces the lift amount. In Embodiment 5 of this invention, it is a flowchart which shows the control performed by ECU. In Embodiment 6 of this invention, it is a block diagram which shows the structure of 1 cylinder of an engine. It is explanatory drawing which shows the injection timing of the fuel in Embodiment 6 of this invention. In Embodiment 6 of this invention, it is explanatory drawing which shows the effect of control using an intake partial cutoff valve. In Embodiment 7 of this invention, it is explanatory drawing which shows the opening / closing timing of an intake valve, and the injection timing of a fuel. In Embodiment 8 of this invention, it is explanatory drawing which shows the fuel-injection timing at the time of starting, etc. In Embodiment 9 of this invention, it is explanatory drawing which shows the valve opening period and fuel injection timing of an intake valve and an exhaust valve. It is explanatory drawing which shows the effect obtained by the injection fuel stabilization control of Embodiment 9. FIG. 38 is a configuration diagram illustrating a modification of the ninth embodiment. In Embodiment 10 of this invention, it is explanatory drawing which shows the valve opening period and fuel injection timing of an intake valve and an exhaust valve. In Embodiment 11 of this invention, it is explanatory drawing which shows the valve opening time of an intake valve, and fuel injection timing. In Embodiment 12 of this invention, it is explanatory drawing which shows the valve opening time of an intake valve, and fuel injection timing. In Embodiment 13 of this invention, it is explanatory drawing which shows the valve opening time of an intake valve, and fuel injection timing. FIG. 5 is a characteristic diagram showing the relationship between the fuel injection amount before and after exhaust top dead center and the engine speed. In Embodiment 14 of this invention, it is a block diagram which expands and shows the structure of 1 cylinder of an engine. In Embodiment 15 of this invention, it is a block diagram which expands and shows the structure of 1 cylinder of an engine. In Embodiment 16 of this invention, it is a block diagram which expands and shows the structure of 1 cylinder of an engine.

Embodiment 1 FIG.
[Configuration of Embodiment 1]
The first embodiment of the present invention will be described below with reference to FIGS. FIG. 1 is an overall configuration diagram for explaining a system configuration according to the first embodiment of the present invention. The system according to the present embodiment includes an engine 10 as an internal combustion engine. Although FIG. 1 illustrates a four-cylinder engine, the present invention is applied to an internal combustion engine having an arbitrary number of cylinders. A combustion chamber 12 is formed in each cylinder of the engine 10 by a piston (not shown), and the piston is connected to a crankshaft that is an output shaft of the engine 10. The engine 10 also includes an intake passage 14 that sucks intake air into each cylinder and an exhaust passage 16 through which exhaust gas is discharged from each cylinder. The intake passage 14 is provided with an electronically controlled throttle valve 18 that adjusts the amount of intake air.

  Each cylinder includes first and second intake ports 20A and 20B and first and second exhaust ports 22A and 22B, respectively. The intake ports 20A and 20B are formed as two parallel passages, and the upstream side is connected to the intake passage 14, and the downstream sides open to the combustion chamber 12 at different positions. Further, the exhaust ports 22A and 22B open to the combustion chamber 12 at different positions on the upstream side, and are connected to the exhaust passage 16 on the downstream side.

  Here, the arrangement of the intake ports 20A and 20B and the exhaust ports 22A and 22B will be described with reference to FIG. FIG. 2 is a configuration diagram showing an enlarged configuration of one cylinder of the engine. As shown in this figure, the intake ports 20A and 20B are arranged on one side and the other side with a reference straight line L as the center. The reference straight line L is defined as a specific straight line extending in the diametrical direction through the center of the circular combustion chamber 12 when the cylinder is viewed from the axial direction (plan view). Similarly, the first and second exhaust ports 22A and 22B are arranged on one side and the other side with the straight line L in between. As a result, the first exhaust port 22A faces the first intake port 20A in a direction parallel to the straight line L, and the second exhaust port 22B also has the second intake port in the direction parallel to the straight line L. Opposite the port 20B.

  Further, as shown in FIG. 1, each cylinder ignites the air-fuel mixture in the combustion chamber 12 (in-cylinder) and the first and second fuel injection valves 24A and 24B disposed in the intake ports 20A and 20B. The spark plug 26 includes first and second intake valves 28A and 28B that open and close the intake ports 20A and 20B, respectively, and first and second exhaust valves 30A and 30B that open and close the exhaust ports 22A and 22B, respectively. ing. The fuel injection valves 24A and 24B are arranged so as to inject fuel toward the openings of the intake ports 20A and 20B.

  For this reason, when the intake valves 28A and 28B are opened, the fuel injected from the fuel injection valves 24A and 24B to the intake ports 20A and 20B directly reaches the cylinder. The fuel injection directions of the fuel injection valves 24A and 24B (centerline of fuel spray spreading in a substantially conical shape) are set in parallel with the straight line L in plan view (see FIG. 3). On the other hand, the spark plug 26 is disposed at the center in the cylinder in plan view.

  The engine 10 also includes an intake variable valve mechanism 32 and an exhaust variable valve mechanism 34. The intake variable valve mechanism 32 changes the valve opening characteristics (opening / closing timing and lift amount) of the intake valves 28A and 28B independently for each valve, and also constitutes a lift amount variable mechanism. Further, the exhaust variable valve mechanism 34 changes the valve opening characteristics of the exhaust valves 30A and 30B independently for each valve. These variable valve mechanisms 32 and 24 are constituted by electromagnetically driven mechanisms as described in, for example, Japanese Patent Application Laid-Open No. 2007-16710.

  That is, the intake variable valve mechanism 32 includes an actuator that opens and closes the intake valve by generating electromagnetic force with a solenoid or the like. Then, the intake variable valve mechanism 32 variably sets the valve opening timing, the valve closing timing, and the lift amount of the intake valves 28A and 28B in accordance with a control signal input to the actuator from the ECU 40 described later. Similarly, the variable exhaust valve mechanism 34 also sets the valve opening characteristics of the exhaust valves 30A and 30B to be variable. The present invention is not limited to the electromagnetically driven variable valve mechanisms 32 and 24. Instead, for example, a swing arm type as described in JP-A-2007-132326 is used. A variable valve mechanism, a VVT (Variable Valve Timing system) as described in JP 2000-87769 A, or the like may be used.

  The engine 10 also includes a swirl control valve (SCV) 36 that constitutes the intake air adjustment valve of the present embodiment. The SCV 36 is composed of, for example, an electronically controlled butterfly valve or the like, and is provided in the intake port 20A on the upstream side of the fuel injection valve 24A. The SCV 36 changes the flow passage area of the intake port 20A according to its opening, and can block the air flow in the intake port 20A in the fully closed position.

  Furthermore, the system of the present embodiment includes a sensor system 38 including various sensors necessary for controlling the engine 10 and a vehicle on which the engine 10 is mounted, and an ECU (Electronic Control Unit) 40 that controls the operating state of the engine 10. I have. The sensor system 38 includes a crank angle sensor that detects a crank angle and an engine speed (engine speed), an air flow sensor that detects an intake air amount, a water temperature sensor that detects the temperature of engine cooling water, and an exhaust air-fuel ratio. An air-fuel ratio sensor, an accelerator opening sensor for detecting an accelerator operation amount (accelerator opening) of the vehicle, and the like are included. The sensor system 38 is connected to the input side of the ECU 40. On the other hand, on the output side of the ECU 40, various actuators including the throttle valve 18, the fuel injection valves 24A and 24B, the spark plug 26, the variable valve mechanisms 32 and 34, the SCV 36, and the like are connected.

  The ECU 40 controls the operation by driving the actuators based on the engine operation information detected by the sensor system 38. Specifically, the engine speed and the crank angle are detected based on the output of the crank angle sensor, and the intake air amount is detected by the air flow sensor. Further, the fuel injection amount is calculated based on the intake air amount, the engine speed, etc., and the fuel injection timing and the ignition timing are determined based on the crank angle. Then, the fuel injection valves 24A and 24B are driven when the fuel injection timing comes, and the spark plug 26 is driven when the ignition timing comes. Further, the ECU 40 drives the variable valve mechanisms 32 and 34 based on the crank angle to open and close the intake valves 28A and 28B and the exhaust valves 30A and 30B at appropriate timing. Thereby, the air-fuel mixture is combusted in the combustion chamber 12 of each cylinder, and the engine 10 can be operated.

[Features of Embodiment 1]
During the operation of the engine, the air flow generated in the intake port changes in a complicated manner depending on the operating state. Accordingly, when fuel is injected into this air flow, the amount of fuel adhering to the wall of the intake port and the intake valve, the adhering position, and the like change irregularly, so that variations in combustion between cycles are likely to occur. For this reason, in the present embodiment, the injected fuel stabilization control described below is executed. In the injected fuel stabilization control, at least during the intake stroke, the SCV 36 is held at the fully closed position, and the air flow in the intake port 20A is shut off. While maintaining this state, the first fuel injection valve 24A is driven during the opening period of the intake valve 28A, and fuel is injected into the intake port 20A where the airflow is blocked.

  At this time, the second fuel injection valve 24B does not execute fuel injection and is held in a stopped state. The intake valves 28A and 28B are opened during a predetermined period in the intake stroke, for example, from the vicinity of the exhaust top dead center to the vicinity of the intake bottom dead center, and the exhaust valves 30A and 30B are closed during the intake stroke. Kept in a state. In the present specification, the “intake stroke” means a period from the exhaust top dead center to the intake bottom dead center, and does not necessarily coincide with the valve opening periods of the intake valves 28A and 28B. In the present specification, the exhaust valves 30A and 30B are opened and closed at a known valve timing unless otherwise specified.

  FIG. 3 is an explanatory diagram showing the effects obtained by the injection fuel stabilization control. According to this control, in the intake stroke, the air flow in the intake port 20A is blocked by the SCV 36, so that the intake air in the intake port 20A stays stably. Then, air is introduced into the cylinder from the intake port 20B. As a result, the fuel injected from the fuel injection valve 24A is not affected by the air flow in the intake port 20A. Therefore, as shown in FIG. 3, while maintaining the natural spray shape, Enter the cylinder from the opening. As a result, the fuel spray shape can be prevented from being disturbed by the air flow, and the amount and position of fuel adhering to the surrounding structure can be stabilized.

  Therefore, according to the present embodiment, it is possible to suppress the variation in combustion between cycles caused by the disturbance of the spray shape, and to expand the operating range in which, for example, lean combustion is possible. Thereby, thermal efficiency can be improved, NOx can be reduced, and exhaust emission can be improved. Further, by shutting off the intake port 20A, as shown in FIG. 3, the intake air flows into the cylinder from the other intake port 20B and forms a swirl flow that circulates along the inner wall surface of the cylinder. By this swirl flow, the homogeneity of the air-fuel mixture is improved and turbulent flow can be generated in the cylinder. Therefore, the combustion rate of the air-fuel mixture can be improved and the thermal efficiency can be further improved.

  In the injection fuel stabilization control described above, the case where the SCV 36 cuts off the air flow in the intake port 20A and performs fuel injection only by the fuel injection valve 24A is exemplified. However, even when the air flow in the intake port 20A is not completely cut off, or when a certain amount of fuel is injected from the second fuel injection valve 24B, the above-described effects can be obtained. Therefore, in the present invention, when the injection fuel stabilization control is performed, the SCV 36 is driven to the closed side within the range where the SCV 36 is not fully closed, and the air flow rate of the intake port 20A is simply reduced compared to the air flow rate of the intake port 20B. Good. Even with this configuration, the air flow in the intake port 20A can be suppressed. In the present invention, the fuel injection amount of the fuel injection valve 24A may be set larger than the fuel injection amount of the fuel injection valve 24B, and then the fuel may be injected by the fuel injection valve 24B. Even with this configuration, it is possible to increase the amount of injected fuel that contributes to suppression of combustion variations.

  Further, in the present invention, without using the SCV 36, any one of the following means (1) to (3) is used to reduce the air flow rate of the intake port 20A from the air flow rate of the intake port 20B. Also good. (1) The shape, structure, passage diameter, etc. of the intake port 20A are configured so that the air is less likely to flow than the intake port 20B. (2) The lift amount of the intake valve 28A is reduced below the lift amount of the intake valve 28B. (3) An air flow rate adjusting mechanism other than SCV, such as a tumble plate, is disposed in the intake port 20A. These means can also suppress the air flow in the intake port 20A.

  On the other hand, when the engine is operated in a high rotation and high load state, the intake air amount and the fuel injection amount are greatly increased. Therefore, it is preferable to open the SCV 36. For this reason, in this Embodiment, it is set as the structure which performs injection fuel stabilization control, when an engine is driving | running states other than high rotation high load. During high rotation and high load operation, with the injected fuel stabilization control stopped, the SCV 36 is opened and air is sucked into the cylinders from both intake ports 20A and 20B, and the fuel injection valves 24A and 24B as required. The fuel is injected by one or both of the above. Thereby, at the time of high rotation and high load, the intake air amount and the fuel injection amount can be sufficiently secured, and the control can be appropriately switched according to the operation state.

  Furthermore, it is preferable to execute the following control during high load operation. That is, in the present embodiment, as the engine load increases, the SCV 36 is driven to the open side and the fuel injection amount of the second fuel injection valve 24B is increased, so that the fuel injection amount of the fuel injection valves 24A and 24B is increased. It is good also as a structure which reduces a difference. According to this configuration, it is possible to sufficiently secure the intake air amount and the fuel injection amount and improve the drivability even during high load operation. This configuration may be executed in combination with Embodiments 2 to 16 described later.

[Specific Processing for Realizing Embodiment 1]
Next, a specific process for realizing the above-described control will be described with reference to FIG. FIG. 4 is a flowchart showing the control executed by the ECU in the first embodiment of the present invention. The routine shown in this figure is repeatedly executed during operation of the engine. In the routine shown in FIG. 4, first, in step 100, it is determined whether or not the engine is operating in a high rotation / high load region. If this determination is not satisfied, the injected fuel stabilization control is executed. That is, in step 102, the SCV 36 is driven to the fully closed position, and the air flow in the intake port 20A is shut off.

  In step 104, the fuel injection valve 24A is driven to cause the fuel injected into the intake port 20A to flow into the cylinder. At this time, the fuel injection timing of the fuel injection valve 24A is set so as to be included in the valve opening period of the intake valve 28A set by the valve timing control. On the other hand, when the determination in step 100 is established, it is during high-speed and high-load operation, so the injected fuel stabilization control is not executed, and in step 106, the SCV 36 is opened. In step 108, fuel is injected by one or both of the fuel injection valves 24A and 24B as necessary.

  In the first embodiment, steps 102 and 104 shown in FIG. 4 show a specific example of the injection state control means in claims 1 and 2, and steps 100, 106, and 108 are high-power control in claim 23. A specific example of the means is shown.

Embodiment 2. FIG.
Next, a second embodiment of the present invention will be described with reference to FIG. The present embodiment is characterized in that the fuel for one cycle is divided into a plurality of times and injected in substantially the same configuration and control (FIGS. 1 and 4) as in the first embodiment. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 2]
FIG. 5 is an explanatory diagram showing split fuel injection in Embodiment 2 of the present invention. As shown in this figure, in the present embodiment, when fuel is injected by the fuel injection valve 24A, the injected fuel for one cycle is divided and injected into a plurality of times. These multiple fuel injections are preferably performed during the opening period of the intake valve 28A. In addition, although the case where the division | segmentation frequency of fuel injection was set to 3 was illustrated in FIG. 5, the frequency | count of division | segmentation injection in this invention is not limited to three times. Further, “TDC” and “BDC” in FIG. 5 indicate an exhaust top dead center and an intake bottom dead center, respectively.

  According to the above control, the injected fuel for one cycle can be divided into several times and injected in small amounts, so that vaporization of the injected fuel can be promoted and the homogeneity of the air-fuel mixture can be improved. Therefore, in addition to the effects of the first embodiment, the combustion speed of the air-fuel mixture can be further improved, the limit of the air-fuel ratio at which lean combustion can be performed (lean limit) can be expanded, and the generation of HC by unburned fuel Can be suppressed. In the second embodiment, FIG. 5 shows a specific example of the divided injection means in claim 3.

Embodiment 3 FIG.
Next, a third embodiment of the present invention will be described with reference to FIG. The present embodiment is characterized in that a phase difference is provided in the closing timing of the first and second intake valves in the configuration and control (FIGS. 1 and 4) substantially the same as in the first embodiment. . In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 3]
FIG. 6 is an explanatory diagram showing opening and closing timings of the first and second intake valves 28A and 28B in the third embodiment of the present invention. As shown in FIG. 6, in the present embodiment, the closing timing of the second intake valve 28B is delayed from the closing timing of the first intake valve 28A. As a specific example, the intake variable valve mechanism 32 sets the closing timing of the intake valve 28A in the vicinity of the intake bottom dead center, and sets the closing timing of the intake valve 28B after the intake bottom dead center. The fuel injection of the fuel injection valve 24A is preferably completed before the intake valve 28A is closed.

  According to the above control, by delaying the closing timing of the intake valve 28B, the intake air amount can be reduced and a so-called Atkinson cycle can be realized. Thereby, in addition to the effect of Embodiment 1, it can improve a thermal efficiency and can improve a fuel consumption. On the other hand, since the closing timing of the intake valve 28A can be advanced, it is possible to prevent the injected fuel from blowing back from the cylinder to the intake port 20A. As a result, the injected fuel that is blown back is not carried over to the next cycle, so the responsiveness during acceleration and deceleration operations can be improved, and combustion variations and torque fluctuations between cycles can be reduced. . In the third embodiment, FIG. 6 shows a specific example of the valve closing timing control means in claim 4.

Embodiment 4 FIG.
Next, a fourth embodiment of the present invention will be described with reference to FIG. The present embodiment is characterized by performing control for protecting the intake valve from deposits in addition to the configuration and control (FIG. 1) substantially the same as in the first embodiment. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 4]
In the injected fuel stabilization control, fuel is mainly injected from the fuel injection valve 24A, so that the intake valve 28A is easily cooled when the injected fuel is vaporized. For this reason, the intake valve 28A tends to be relatively low in temperature, and there is a problem that deposits are likely to adhere to the intake valve 28A whose temperature has decreased. Therefore, in the present embodiment, the deposit suppression control described below is intermittently executed to protect the intake valve 28A from the deposit.

(Deposit suppression control)
This control is executed in place of the injected fuel stabilization control when the exhaust air-fuel ratio is close to the stoichiometric air-fuel ratio (stoichiometric), stops the fuel injection by the first fuel injection valve 24A, The fuel is injected from the second fuel injection valve 24B. According to the deposit suppression control, since fuel is not injected in the vicinity of the intake valve 28A, the temperature of the intake valve 28A can be raised, and deposits can be suppressed from adhering to the intake valve. If the fuel injection amount is constant, the stoichiometric amount of air in the cylinder is small compared to the lean state and the heat capacity of the air is small, so the combustion temperature tends to increase by that amount. Therefore, if the deposit suppression control is executed with stoichiometry, the intake valve 28A can be efficiently heated in a short time.

  On the other hand, when the air-fuel ratio is other than stoichiometric (practically lean combustion is performed, the air-fuel ratio is leaner than the stoichiometric air-fuel ratio), the above-described injected fuel stabilization control is executed. Therefore, according to the present embodiment, the valve can be heated using the timing at which the air-fuel ratio becomes stoichiometric while performing the injected fuel stabilization control during lean combustion. Thereby, in addition to the effect of Embodiment 1, the deposit suppression effect can be acquired.

  In the above description, the case where the injected fuel stabilization control and the deposit suppression control are switched based on the air-fuel ratio is exemplified. However, even when the deposit suppression control is executed in a situation other than stoichiometry, the above-described effects can be obtained. Therefore, in the present invention, it is not always necessary to switch the control based on the air-fuel ratio. For example, the deposit suppression control may be executed at a constant cycle while executing the injected fuel stabilization control. Further, for example, the temperature of the intake valve may be estimated based on the operating state of the engine, and the deposit suppression control may be executed when the estimated temperature is equal to or lower than the allowable temperature.

  Further, in the present invention, when the deposit suppression control is executed, it is not always necessary to stop the fuel injection of the fuel injection valve 24A, and only the fuel injection amount of the fuel injection valve 24A may be reduced. That is, in the deposit suppression control, the fuel injection amount of the fuel injection valve 24A is set larger than the fuel injection amount of the fuel injection valve 24A, and the fuel injection amount of the fuel injection valve 24A is set in a state where the entire fuel injection amount is constant. It is good also as a structure decreased relatively. Even with this configuration, the temperature of the intake valve 28A can be increased.

[Specific processing for realizing Embodiment 4]
Next, specific processing for realizing the above-described control will be described with reference to FIG. FIG. 7 is a flowchart showing the control executed by the ECU in the fourth embodiment of the present invention. Since the routine shown in this figure is obtained by adding steps 202 and 204 to the first embodiment (FIG. 4), description other than the added steps is omitted. In the routine shown in FIG. 7, first, in step 202, it is determined whether or not the air-fuel ratio detected by the air-fuel ratio sensor or the like is stoichiometric. If this determination is established, in step 204, the above-described deposit suppression control is executed. On the other hand, if the determination in step 202 is not established, injected fuel stabilization control is executed in steps 206 and 208.

  In the fourth embodiment, step 202 shown in FIG. 7 shows a specific example of the control switching means in claims 5 and 6, and step 204 shows a specific example of the injection state changing means. Steps 206 and 208 show specific examples of the injection state control means in claims 1 and 2, and steps 200, 210 and 212 show specific examples of the high output time control means in claim 23.

Embodiment 5 FIG.
Next, a fifth embodiment of the present invention will be described with reference to FIGS. The present embodiment is characterized in that, in addition to the configuration and control (FIGS. 1 and 4) substantially the same as those in the first embodiment, the lift amount of the second intake valve is reduced when the engine is started. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 5]
When the engine is started, the temperature inside the cylinder is low, so that the injected fuel is not sufficiently vaporized, and the injected fuel tends to adhere to the inner wall surface of the cylinder. For this reason, in the present embodiment, when the engine is started, the lift amount of the second intake valve 28B is reduced by the intake variable valve mechanism 32 as compared with the lift amount after the start. Here, “at the time of starting” corresponds to a period from when the engine is started to when warm-up is completed, for example. Further, the above-described lift amount control at the start may be performed only when the temperature of the engine coolant is in a low temperature region corresponding to the cold start, for example.

  FIG. 8 is an explanatory diagram showing the operational effect of the control for reducing the lift amount in the fifth embodiment of the present invention. As shown in this figure, according to the present embodiment, at the time of starting, the flow area of the intake port 20B is decreased by the intake valve 28B, and the flow velocity of the air flowing into the cylinder from the intake port 20B can be increased. it can. As a result, the swirl flow in the cylinder can be strengthened in the direction opposite to the fuel injection direction, the momentum of the injected fuel can be reduced by this swirl flow, and the injected fuel can be diffused efficiently. As a result, vaporization of the injected fuel can be promoted even during a cold start, and the fuel can be prevented from adhering to the cylinder inner wall surface. Therefore, in addition to the effects of the first embodiment, the engine startability and the exhaust emission at the start can be improved.

[Specific Processing for Realizing Embodiment 5]
Next, specific processing for realizing the above-described control will be described with reference to FIG. FIG. 9 is a flowchart showing the control executed by the ECU in the fifth embodiment of the present invention. The routine shown in this figure is repeatedly executed during engine operation in parallel with the routine shown in the first embodiment (FIG. 4). In the routine shown in FIG. 9, first, in step 300, it is determined whether or not the engine is being started. If this determination is established, in step 302, the lift amount of the intake valve 28B is decreased to a predetermined start-time lift amount by the intake variable valve mechanism 32. The starting lift amount is set in advance as, for example, a lift amount that can sufficiently strengthen the swirl flow in the cylinder. On the other hand, if the determination in step 300 is not satisfied, in step 304, the lift amount of the intake valve 28B is returned to the lift amount at the normal time (after completion of warm-up).

  In the fifth embodiment, steps 300 and 302 in FIG. 9 show a specific example of the starting lift amount reducing means in claim 7. In the fifth embodiment, the control has been described on the premise of the first embodiment. However, the present invention is not limited to this, and for example, the configuration adopted in any one of the second, third, and fourth embodiments and the embodiment. You may combine the structure of form 5 in the realizable range.

Embodiment 6 FIG.
Next, a sixth embodiment of the present invention will be described with reference to FIGS. The present embodiment is characterized in that, in addition to the configuration and control (FIGS. 1 and 4) substantially the same as those of the first embodiment, an intake partial shutoff valve is disposed in the second intake port. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 6]
FIG. 10 is a configuration diagram showing the configuration of one cylinder of the engine in the sixth embodiment of the present invention. As shown in this figure, in the present embodiment, an intake partial shutoff valve 50 is arranged in the second intake port 20B. The intake partial shutoff valve 50 is configured by, for example, an electronically controlled butterfly valve, for example, similar to the SCV 36, and can close the intake port 20B on the upstream side of the fuel injection valve 24B. However, a notch 52 is provided in a portion of the intake partial shut-off valve 50 that allows air to flow into the outer periphery of the cylinder, that is, a portion away from the center (or straight line L) of the combustion chamber 12. As a result, in the state where the intake partial shutoff valve 50 is driven to the fully closed position, the intake port 20B is blocked except for the position of the notch 52, and the air in the intake port 20B passes through the notch 52 in the cylinder outer periphery. It is comprised so that it may flow into.

  FIG. 11 is an explanatory diagram showing the fuel injection timing in the sixth embodiment. In the present embodiment, fuel is injected into the first half of the intake stroke by the first fuel injection valve 24A while the intake partial cutoff valve 50 is driven to the closed side (preferably, the fully closed position) during the intake stroke. In the present invention, the intake partial cutoff valve 50 may be driven to close at least in the second half of the intake stroke, and the intake partial cutoff valve 50 may be opened in the first half of the intake stroke. Further, in the present embodiment, the valve opening periods of the intake valves 28A and 28B are set so as to cover almost the entire intake stroke, for example.

  FIG. 12 is an explanatory diagram showing the operational effect of the control using the intake partial cutoff valve. According to the above control, first, in the first half of the intake stroke, an air-fuel mixture is formed in the cylinder by the fuel injected from the fuel injection valve 24A and the air flowing in from the intake port 20B. Next, in the second half of the intake stroke, fuel injection is completed, and only air flows from the notch 52 into the cylinder outer periphery. This air surrounds the air-fuel mixture formed in the first half of the intake stroke while flowing along the inner wall surface of the cylinder. As a result, at the end of the intake stroke, as shown in FIG. 12, an air-fuel mixture layer is formed at the center of the cylinder, and an air layer is formed at the outer periphery of the cylinder, so that the gas in the cylinder is stratified. The

  Therefore, according to the present embodiment, in addition to the effects of the first embodiment, the air-fuel mixture can be concentrated in the vicinity of the spark plug 26 and stratified combustion can be realized. Thereby, even with a small amount of fuel injection, combustion can be stabilized and lean combustion can be easily performed. Moreover, pump loss and cooling loss can be suppressed, and fuel consumption can be improved. In the sixth embodiment, FIG. 11 shows a specific example of the injection state control means in claim 8. In the sixth embodiment, the control has been described on the premise of the first embodiment. However, the present invention is not limited to this, and for example, the configuration adopted in any one of the second to fifth embodiments and the sixth embodiment. These configurations may be combined in a realizable range.

Embodiment 7 FIG.
Next, a seventh embodiment of the present invention will be described with reference to FIG. The present embodiment is characterized in that, in addition to the configuration and control of the sixth embodiment, a phase difference is provided in the closing timing of the first and second intake valves as in the third embodiment. Yes. In the present embodiment, the same components as those in the sixth embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 7]
FIG. 13 is an explanatory diagram showing the opening / closing timing of the intake valve and the fuel injection timing in Embodiment 7 of the present invention. In the present embodiment, as in the sixth embodiment, fuel is injected into the first half of the intake stroke by the fuel injection valve 24A while the intake partial cutoff valve 50 is closed during the intake stroke. As in the third embodiment, the closing timing of the intake valve 28B is delayed from the closing timing of the intake valve 28A.

  According to the above control, in addition to the operational effects of the sixth embodiment, the operational effects of the Atkinson cycle described in the third embodiment can be obtained. Moreover, in the above-described Atkinson cycle, the air-fuel mixture in the cylinder may blow back to the intake port 20B due to the late timing of the intake valve 28B. On the other hand, in the present embodiment, even if the blow-back to the intake port 20B occurs, the blow-back gas becomes an air layer on the outer periphery in the cylinder. Therefore, it is possible to suppress the air-fuel mixture amount in the cylinder from fluctuating due to blow-back and reduce the combustion variation between cycles. In addition, in the said Embodiment 7, FIG. 13 has shown the specific example of the valve closing timing control means in Claim 9. FIG.

Embodiment 8 FIG.
Next, an eighth embodiment of the present invention will be described with reference to FIG. In addition to the configuration and control of the sixth embodiment, the present embodiment is characterized in that fuel is injected even in the latter half of the intake stroke at the time of starting. In the present embodiment, the same components as those in the sixth embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 8]
FIG. 14 is an explanatory diagram showing the fuel injection timing and the like at the start in the eighth embodiment of the present invention. In the present embodiment, when the engine is started, fuel is injected into the first half and the second half of the intake stroke by the fuel injection valve 24A while the intake partial cutoff valve 50 is closed during the intake stroke. Further, at the time of starting, the ignition timing is retarded from the compression top dead center.

  According to the above control, the fuel injection in the first half of the intake stroke forms an air-fuel mixture layer in the vicinity of the spark plug 26 as described in the sixth embodiment, so that the ignitability at the start is improved by this air-fuel mixture layer. be able to. Further, the fuel injection in the latter half of the intake stroke can form an air-fuel mixture also on the inner and outer periphery of the cylinder, and can increase the combustion amount of the air-fuel mixture and raise the exhaust gas temperature. Furthermore, the exhaust gas temperature can be increased by retarding the ignition timing. Therefore, according to the present embodiment, in addition to the operational effects of the sixth embodiment, the engine startability can be enhanced, and the catalyst can be quickly warmed up even during a cold start.

  In the above control, the fuel injection in the latter half of the intake stroke is to inject auxiliary fuel for forming an air-fuel mixture also in the cylinder outer periphery. For this reason, in the present embodiment, the fuel injection amount in the latter half of the intake stroke may be configured to be smaller than the fuel injection amount in the first half of the intake stroke. In the eighth embodiment, the control is described on the premise of the sixth embodiment. However, the present invention is not limited to this, and for example, the configuration adopted in the seventh embodiment and the configuration of the eighth embodiment can be realized. You may combine in the range. In the eighth embodiment, FIG. 14 shows a specific example of the injection state control means and the ignition retard means in claim 10.

Embodiment 9 FIG.
Next, a ninth embodiment of the present invention will be described with reference to FIGS. In the present embodiment, in addition to the configuration and control (FIGS. 1 and 4) of the first embodiment, the exhaust valve is opened during the intake stroke. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 9]
FIG. 15 is an explanatory diagram showing the valve opening periods and fuel injection timings of the intake and exhaust valves in the ninth embodiment of the present invention. As shown in this figure, in the present embodiment, in the first half of the intake stroke, the first and second intake valves 28A and 28B are opened, and the first and second exhaust valves 30A and 30B are closed. . In the second half of the intake stroke, the intake valves 28A and 28B and the second exhaust valve 30B are closed, and only the first exhaust valve 30A is opened. Further, in the present embodiment, the intake partial cutoff valve 50 is driven to the closed side during the intake stroke, and fuel is injected into the first half of the intake stroke by the fuel injection valve 24A. In a stroke other than the intake stroke, each valve 28A, 28B, 30A, 30B is opened and closed at a known timing.

  FIG. 16 is an explanatory diagram showing the effects obtained by the injected fuel stabilization control of the ninth embodiment. According to this control, first, in the first half of the intake stroke, almost in the same way as in the first embodiment, the air-fuel mixture is in the cylinder by the fuel injected from the fuel injection valve 24A and the air flowing in from the intake port 20B. It is formed. Next, in the second half of the intake stroke, the exhaust valve 30A is opened with the intake valves 28A and 28B closed, so that the exhaust gas in the exhaust port 22A is introduced into the cylinder outer periphery by the lowering operation of the piston. The Thereby, at the end of the intake stroke, as shown in FIG. 16, the mixture layer is formed in the center of the cylinder and the exhaust gas (EGR gas) layer is formed on the outer periphery of the cylinder, that is, the stratified EGR is Can be realized.

  Here, the generation process of the stratified EGR will be described in detail. First, in the first half of the intake stroke, air flows from the second intake port 20B, and therefore, in the cylinder, as shown in FIG. A swirl flow (for example, counterclockwise) is generated. In the latter half of the intake stroke, only the exhaust port 22A that is not opposed to the intake port 20B into which air is introduced is opened among the exhaust ports 22A and 22B. As a result, the exhaust gas flows from the exhaust port 22A into the cylinder outer periphery along the swirl flow direction. Therefore, the exhaust gas can be smoothly introduced into the outer periphery of the cylinder using the swirl flow, and the stratified EGR can be realized.

  As described above, according to the present embodiment, the gas in the cylinder can be stratified, and in addition to the operational effects of the first embodiment, the fuel efficiency can be improved by the same principle as in the sixth embodiment. Can do. In addition, since an EGR gas layer can be formed on the outer periphery of the cylinder, EGR can be executed, the air-fuel ratio can be maintained at the stoichiometric air-fuel ratio, and NOx in the exhaust gas can be reduced by the three-way catalyst. In the ninth embodiment, FIG. 15 shows a specific example of the injection state control means and the valve opening period control means in claim 11. In the ninth embodiment, the control has been described on the premise of the first embodiment. However, the present invention is not limited to this, and for example, the configuration adopted in any of the second to eighth embodiments and the ninth embodiment. These configurations may be combined in a realizable range.

  On the other hand, in the ninth embodiment, a configuration as in the modification shown in FIG. 17 may be adopted. FIG. 17 is a configuration diagram showing a modification of the ninth embodiment. As shown in this figure, in the ninth embodiment, the intake partial shutoff valve 50 is arranged in the second intake port 20B, and the exhaust partial shutoff valve 60 is arranged in the first exhaust port 22A. Here, the intake partial shutoff valve 50 is the same as that described in the sixth embodiment. Further, the exhaust partial shutoff valve 60 is configured by a butterfly valve or the like that is substantially the same as the intake partial shutoff valve 50, and the exhaust port 22A can be closed except for a portion corresponding to the inner and outer circumferences of the exhaust port 22A. It has become. That is, a notch 62 is provided in a portion of the exhaust partial shutoff valve 60 that is away from the center (or straight line L) of the combustion chamber 12. The exhaust partial shutoff valve 60 is driven to the closing side (preferably, the fully closed position) in the second half of the intake stroke.

  According to the configuration of the modified example, the intake partial shutoff valve 50 can strengthen the swirl flow in the cylinder in the first half of the intake stroke, and can bring the position where the swirl flow flows closer to the cylinder inner wall surface. it can. As a result, in the second half of the intake stroke, exhaust gas can be efficiently guided to the outer periphery of the cylinder by a strong swirl flow, and turbulence of the stratified EGR can be reduced. On the other hand, the exhaust partial shutoff valve 60 can force exhaust gas to flow into the cylinder outer periphery in the latter half of the intake stroke, and can prevent the exhaust gas from flowing into the cylinder center. Therefore, the stratified EGR can be more stably realized by any of the shutoff valves 50 and 60. In addition, although the case where both the shut-off valves 50 and 60 were used was illustrated in the above modification, the present invention is not limited to this, and only one of the shut-off valves 50 and 60 may be used.

Embodiment 10 FIG.
Next, Embodiment 10 of the present invention will be described with reference to FIG. In the present embodiment, in the configuration and control of the ninth embodiment, the first and second intake valves are opened during the first half and the second half of the intake stroke, and fuel injection is performed in the first half and the second half of the intake stroke, respectively. It is characterized by. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 10]
FIG. 18 is an explanatory diagram showing the valve opening periods and fuel injection timings of the intake and exhaust valves in the tenth embodiment of the present invention. As shown in this figure, in the present embodiment, the intake valves 28A and 28B are opened over the first half and the second half of the intake stroke. As for the exhaust valves 30A and 30B, as in the ninth embodiment, both the exhaust valves 30A and 30B are closed in the first half of the intake stroke, and only the first exhaust valve 30A is opened in the second half of the intake stroke. To do. Further, fuel injection by the fuel injection valve 24A is executed in the first half and the second half of the intake stroke.

  According to the above control, as described in the ninth embodiment, an air-fuel mixture can be formed in the cylinder in the first half of the intake stroke, and exhaust gas can be introduced into the outer periphery of the cylinder in the second half of the intake stroke. A stratified EGR can be realized. On the other hand, there is a problem that HC (hydrocarbon) in the exhaust gas tends to increase due to unburned fuel remaining on the inner and outer periphery of the cylinder because the gas on the inner and outer periphery of the cylinder is difficult to burn when performing the stratified EGR. On the other hand, in the present embodiment, since the intake valve 28B is opened also in the latter half of the intake stroke, air is also supplied to the exhaust gas layer on the inner and outer periphery of the cylinder. Fuel injected in the second half is also supplied. Thereby, a gas layer in which exhaust gas and air-fuel mixture are mixed is formed on the outer periphery in the cylinder.

  This gas layer is difficult to burn by flame propagation from the central part of the cylinder, but burns by self-ignition by the heat of exhaust gas contained in itself in the latter half of the combustion stroke. Therefore, according to the present embodiment, it is possible to reliably burn the unburned fuel that tends to remain on the outer periphery of the cylinder by using the heat of the exhaust gas introduced into the cylinder. Thereby, in addition to the operational effects of the ninth embodiment, HC in the exhaust gas can be suppressed and exhaust emission can be improved.

  In the above control, for the same reason as in the eighth embodiment, the fuel injection amount in the latter half of the intake stroke may be made smaller than the fuel injection amount in the first half of the intake stroke. Further, in the present invention, for example, the configuration employed in any of Embodiments 2 to 8 and the configuration of Embodiment 10 may be combined within a realizable range. Furthermore, in this invention, it is good also as a structure which applies any one or both of the shut-off valves 50 and 60 employ | adopted by the modification of Embodiment 9 to Embodiment 10. FIG. Thereby, stratification EGR can be realized more stably. In the tenth embodiment, FIG. 18 shows a specific example of the injection state control means and the valve opening period control means in claim 12.

Embodiment 11 FIG.
Next, an eleventh embodiment of the present invention will be described with reference to FIG. In the present embodiment, in addition to the configuration and control (FIGS. 1 and 4) of the first embodiment, the first intake valve is opened until the second intake valve is opened. It is characterized by injecting fuel. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 11]
FIG. 19 is an explanatory diagram showing the valve opening timing and the fuel injection timing of the intake valve in the eleventh embodiment of the present invention. As shown in this figure, in the present embodiment, the opening timing of the first intake valve 28A is made later than the exhaust top dead center (TDC), and the opening timing of the second intake valve 28B is set to the first timing. 1 is later than the opening timing of the intake valve 28A. Then, the fuel is injected by the first fuel injection valve 24A between the time when the intake valve 28A is opened and the time when the intake valve 28B is opened. The exhaust valves 30A and 30B are kept closed during the opening period of the intake valves 28A and 28B.

  According to the above control, when the fuel is injected, the intake port 20A is blocked by the SCV 36, and the other ports 20B, 22A, and 22B are closed by the valves 28B, 30A, and 30B, respectively, and the piston performs the lowering operation. become. As a result, fuel can be injected into the cylinder in a negative pressure state, so that the boiling point of the fuel can be lowered and vaporization of the injected fuel can be promoted. And a favorable air-fuel mixture can be formed quickly, and the injected fuel can be prevented from adhering to the inner wall surface of the cylinder. Therefore, if the above control is executed when the engine is started, in addition to the operational effects of the first embodiment, the startability at the cold start and the exhaust emission can be improved.

  As described above, the control according to the present embodiment has a remarkable effect at the time of starting the engine. Therefore, the control may be executed only at the time of starting (particularly at the time of starting at a low temperature). On the other hand, the above control can achieve a sufficient effect by promoting the vaporization of the injected fuel even in an operating state other than at the time of starting. Therefore, in the present invention, in any operating state including after starting (after completion of warm-up). The above control may be executed. Further, in the present invention, for example, the configuration adopted in any of Embodiments 2 to 10 and the configuration of Embodiment 11 may be combined in a realizable range. In the eleventh embodiment, FIG. 19 shows a specific example of the injection state control means and the valve opening timing control means in claim 15.

Embodiment 12 FIG.
Next, Embodiment 12 of the present invention will be described with reference to FIG. In the present embodiment, fuel is injected before exhaust top dead center when performing auto-ignition combustion control on the premise of the configuration and control (FIGS. 1 and 4) substantially the same as those of the first embodiment. It is characterized by. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 12]
In the present embodiment, the self-ignition combustion control is performed to self-ignite the air-fuel mixture in the cylinder by compression. The self-ignition combustion control is a known technique described in, for example, Japanese Patent Laid-Open No. 2006-2638. More specifically, in the self-ignition combustion control, the air-fuel mixture is spontaneously ignited basically without using a spark plug by bringing the air-fuel mixture in the cylinder into a high temperature / high pressure state in the compression stroke. According to this control, homogeneous combustion occurs in the entire cylinder, so that the air-fuel mixture can be made lean, but there is a problem that misfire is likely to occur in the low load operation region. For this reason, the present embodiment aims to reform the injected fuel so that it is easily ignited (so-called fuel reforming) before the start of the compression stroke.

  FIG. 20 is an explanatory diagram showing the valve opening timing and fuel injection timing of the intake valve in the twelfth embodiment of the present invention. As shown in this figure, in this embodiment, the exhaust valves 30A and 30B are closed before the exhaust top dead center (TDC), and the exhaust top dead center arrives after the exhaust valves 30A and 30B are closed. In the meantime, the first intake valve 28A is opened. At this time, the second intake valve 28B opens at least after exhaust top dead center (preferably after the intake valve 28A is closed). Then, fuel is injected by the first fuel injection valve 24A from when the first intake valve 28A is opened until when exhaust top dead center arrives. The SCV 36 is closed to at least the intake bottom dead center (BDC) after the intake valve 28A is opened.

  According to the above control, during the period from when the first intake valve 28A is opened until the second intake valve 28B is opened, the intake port 20A is blocked by the SCV 36, and the other ports 20B, 22A. , 22B are closed by valves 28B, 30A, 30B, respectively. As a result, the inside of the cylinder is hermetically sealed, so that the exhaust gas remaining in the cylinder is once compressed before the exhaust top dead center and becomes high temperature. When the fuel is injected in this state, the injected fuel is reformed into a radical state or component that easily ignites in high-temperature exhaust gas. Then, as shown in FIG. 20, the reformed fuel is easily self-ignited and burned in the compression stroke.

  Therefore, according to the present embodiment, fuel reforming can be performed by injecting fuel into the high-temperature exhaust gas once compressed in the cylinder. Thereby, at the time of execution of the self-ignition combustion control, the injected fuel can be stably self-ignited even in the low load operation region, and misfires can be prevented. In the present embodiment, the control shown in FIG. 20 is executed on the premise of self-ignition combustion control. However, the present invention is not limited to the self-ignition combustion control, and is also applied, for example, when the self-ignition combustion control is not executed (when the air-fuel mixture is ignited by the spark plug). Also in this case, the injected fuel can be easily burned by the fuel reforming effect, and the ignitability of the air-fuel mixture can be improved. Moreover, vaporization of the injected fuel can be promoted by the high-temperature exhaust gas, and the fuel can be prevented from adhering to the inner wall surface of the cylinder.

  Further, in the present invention, for example, the configuration employed in any of Embodiments 2 to 11 and the configuration of Embodiment 12 may be combined within a realizable range. In the twelfth embodiment, FIG. 20 shows a specific example of the injection state control means and the opening / closing timing control means in claim 16, and shows a specific example of the self-ignition combustion control means in claim 18. .

Embodiment 13 FIG.
Next, a thirteenth embodiment of the present invention will be described with reference to FIGS. In the present embodiment, in the twelfth embodiment, fuel is injected before and after exhaust top dead center, and the fuel injection amount at the time of these injections is variably set according to the engine speed. It is a feature. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 13]
FIG. 21 is an explanatory diagram showing the valve opening timing and fuel injection timing of the intake valve in the thirteenth embodiment of the present invention. FIG. 22 is a characteristic diagram showing the relationship between the fuel injection amount before and after exhaust top dead center and the engine speed. In the present embodiment, as shown in FIG. 21, after performing the same control as in the twelfth embodiment (FIG. 20), the first fuel injection valve 24A supplies fuel before and after exhaust top dead center. Spray. As shown in FIG. 22, when the engine speed is equal to or higher than a predetermined reference value, the fuel injection amount before the exhaust top dead center (TDC) is increased more than the fuel injection amount after the exhaust top dead center. Further, when the engine speed is less than the reference value, the fuel injection amount after the exhaust top dead center is increased more than the fuel injection amount before the exhaust top dead center.

  This embodiment is premised on performing self-ignition combustion control. The fuel injection before the exhaust top dead center compresses the injected fuel together with the exhaust gas in the cylinder and can perform the fuel reforming over a relatively long time, so that the maximum fuel reforming effect can be obtained. it can. In addition, the fuel injection after exhaust top dead center is less in fuel reforming effect than the fuel injection before exhaust top dead center, but at a timing close to the valve opening timing of the intake valve 28B (air introduction timing). Since the fuel is injected, an air-fuel mixture with high homogeneity can be formed.

  When the engine speed is high, the time from fuel injection to ignition (that is, the time during which the air-fuel mixture is formed) is short. Therefore, in order to obtain a sufficient fuel reforming effect, the fuel before exhaust top dead center is required. It is preferable to increase the injection amount. For this reason, in the present embodiment, as shown in FIG. 22, the higher the engine speed, the more the fuel injection amount before the exhaust top dead center is increased. The fuel injection amount before the exhaust top dead center is set larger than that after the exhaust top dead center. FIG. 22 shows the fuel injection amounts before and after exhaust top dead center on the assumption that the total amount of fuel injection is constant.

  On the other hand, when the engine speed is low, the time from fuel injection to ignition is relatively long, so the fuel injection amount after exhaust top dead center is increased to achieve both the fuel reforming effect and the homogeneity of the mixture. Is preferred. For this reason, in this embodiment, the lower the engine speed, the greater the fuel injection amount after exhaust top dead center. In particular, when the engine speed is less than the reference value, the fuel after exhaust top dead center is increased. Set the injection amount higher than before exhaust top dead center. Note that the reference value corresponds to the minimum engine speed at which it is preferable to increase the fuel injection amount before the exhaust top dead center, for example, from the viewpoint of the fuel reforming level.

  According to the above control, the fuel injection amount before the exhaust top dead center can be increased at high revolutions, and a sufficient fuel reforming effect can be obtained even if the time from fuel injection to ignition is short. Moreover, at the time of low rotation, the fuel injection amount after exhaust top dead center can be increased, and both the fuel reforming effect and the homogeneity of the air-fuel mixture can be achieved. Therefore, the fuel reforming level can be appropriately controlled according to the engine speed.

  In the thirteenth embodiment, FIG. 21 shows a specific example of the means for injecting around the top dead center in claim 17, and FIG. 22 shows a specific example of the injection amount control means. In the present invention, the configuration employed in any of Embodiments 2 to 11, for example, may be combined with the configuration of Embodiment 13 within a feasible range.

Embodiment 14 FIG.
Next, a fourteenth embodiment of the present invention will be described with reference to FIG. The present embodiment is characterized in that, in the configuration and control (FIGS. 1 and 4) of the first embodiment, the fuel injection direction of the first fuel injection valve is inclined toward the center side of the combustion chamber. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 14]
FIG. 23 is an enlarged configuration diagram showing the configuration of one cylinder of the engine in the fourteenth embodiment of the present invention. As shown in this figure, in the present embodiment, the fuel injection direction of the first fuel injection valve 70 (the center line L0 of the fuel spray spreading in a substantially conical shape) is set to the combustion chamber with respect to the reference straight line L. It is inclined toward the center O of 12. That is, the fuel injection valve 70 is arranged so that the spray center line L 0 passes through the center O of the combustion chamber 12.

  According to the above configuration, the fuel injection valve 70 can inject fuel toward the center O of the combustion chamber 12. Thereby, as shown in FIG. 23, most of the injected fuel does not collide with the swirl flow that flows along the inner periphery of the cylinder from the front, so that it is possible to prevent the swirl flow from being weakened by the spray of fuel. it can. Therefore, in addition to the operational effects of the first embodiment, the swirl flow can be maintained until immediately before combustion. Therefore, the homogenization of the air-fuel mixture is promoted by the turbulent air flow, and the combustibility and fuel consumption can be improved. it can. Further, the injected fuel can be injected in the diameter direction of the combustion chamber 12, that is, the direction in which the distance to the cylinder inner wall surface becomes the longest, and the fuel can be prevented from adhering to the cylinder inner wall surface.

Embodiment 15 FIG.
Next, Embodiment 15 of the present invention will be described with reference to FIG. The present embodiment is characterized in that the fuel injection angle of the first fuel injection valve is set wide on the premise of the configuration of the fourteenth embodiment. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 15]
FIG. 24 is an enlarged configuration diagram showing the configuration of one cylinder of the engine in the fifteenth embodiment of the present invention. As shown in this figure, the fuel injection angle θ of the first fuel injection valve 80 is such that the center O1 of the second intake valve 28B and the center O2 of the first exhaust valve 30A are inside the fuel injection angle θ. It is set to an angle that fits. Here, the fuel injection angle θ is defined as an angle at which the fuel injected from the fuel injection valve 80 in a substantially conical shape expands in a plan view. Further, the center line L0 of the fuel injection angle θ is arranged so as to pass through the center O of the combustion chamber 12 as in the fourteenth embodiment. Further, the centers O1, O2 of the valves 28B, 30A correspond to, for example, positions where the valve stems are arranged.

  According to the above configuration, most of the in-cylinder space can be accommodated within the range of the fuel injection angle θ. Thereby, the injected fuel can be dispersed in a wide range in the cylinder, fuel vaporization can be promoted, and a homogeneous air-fuel mixture can be formed throughout the cylinder.

Embodiment 16 FIG.
Next, a sixteenth embodiment of the present invention will be described with reference to FIG. In the present embodiment, in the configuration and control (FIGS. 1 and 4) of the first embodiment, the center line of the fuel spray injected from the first fuel injection valve is set to be second than the center of the combustion chamber. It is characterized in that it is inclined toward the fuel injection valve side. In the present embodiment, the same components as those in the first embodiment are denoted by the same reference numerals, and the description thereof is omitted.

[Features of Embodiment 16]
FIG. 25 is an enlarged configuration diagram showing the configuration of one cylinder of the engine in the sixteenth embodiment of the present invention. As shown in this figure, in the present embodiment, the center line L0 of the fuel spray injected from the first fuel injection valve 90 is set to be a second fuel injection valve from the center O of the combustion chamber 12 in plan view. It arrange | positions so that it may pass through the position near 24B. According to this configuration, as shown in FIG. 25, the direction of fuel injection can be matched to the direction of the swirl flow that flows along the outer periphery of the cylinder, and the swirl flow can be strengthened by fuel spraying. Thereby, in addition to the effect of Embodiment 1, a swirl flow can be hold | maintained just before combustion. Therefore, the homogenization of the air-fuel mixture can be promoted by the turbulent air flow, and the combustibility and fuel consumption can be improved.

  In the above configuration, for example, as shown in FIG. 11 used in Embodiment 6, the fuel injection period of the first fuel injection valve 24A is set to the first half of the intake stroke, and the fuel injection is performed in the first half of the intake stroke. Preferably completed. Accordingly, the swirl flow can be accelerated by the fuel spray at the first half of the intake stroke when the swirl flow generated in the cylinder is relatively weak. Therefore, the swirl flow in the cylinder can be strengthened from the first half of the intake stroke, and the effects of the sixteenth embodiment can be exhibited more remarkably. In addition, when fuel is injected into a strong swirl flow, the fuel droplets are likely to adhere to the inner wall surface of the cylinder due to centrifugal force before vaporizing the fuel. Therefore, by injecting the fuel when the swirl flow is weak, it is possible to prevent the injected fuel from adhering to the inner wall surface of the cylinder.

  In addition, Embodiments 14 to 16 have been described on the premise of the configuration and control of Embodiment 1. However, the present invention is not limited to this, and for example, the configuration employed in any of Embodiments 2 to 13 may be combined with the configuration of Embodiments 14 to 16 within a possible range.

10 Engine (Internal combustion engine)
12 Combustion chamber 14 Intake passage 16 Exhaust passage 18 Throttle valves 20A, 20B First and second intake ports 22A, 22B First and second exhaust ports 24A, 70, 80, 90 First fuel injection valve 24B Second Fuel injection valve 26 Spark plugs 28A and 28B First and second intake valves 30A and 30B First and second exhaust valves 32 Intake variable valve mechanism (lift amount variable mechanism)
34 Exhaust variable valve mechanism 36 Swirl control valve (intake adjustment valve)
38 Sensor system 40 ECU
50 Intake partial shutoff valve 60 Exhaust partial shutoff valve 52, 62 Notch

Claims (20)

First and second intake ports that are connected to the intake passage on the upstream side and open into the cylinder at different positions on the downstream side;
First and second intake valves provided in the first and second intake ports;
First and second fuel injection valves for injecting fuel into the first and second intake ports;
An intake adjustment valve that is provided in the first intake port upstream of the first fuel injection valve and changes a flow passage area of the first intake port;
At least in the intake stroke, the intake adjustment valve is driven to the closed side to reduce the air flow rate of the first intake port to be lower than the air flow rate of the second intake port, and the first fuel injection valve Injection state control means for executing fuel injection in a state in which the fuel injection amount is set larger than the fuel injection amount of the second fuel injection valve;
A shutoff valve capable of closing the intake port except for a portion of the second intake port that allows air to flow into the outer periphery of the cylinder, wherein the intake portion is cut off at least in the second half of the intake stroke. A valve,
The control apparatus for an internal combustion engine, wherein the injection state control means is configured to inject fuel in the first half of an intake stroke by the first fuel injection valve.   The injection state control means closes the intake adjustment valve and shuts off the air flow in the first intake port, and controls the first fuel injection valve during the opening period of the first intake valve. The control device for an internal combustion engine according to claim 1, wherein the control device is configured to inject fuel.   The control apparatus for an internal combustion engine according to claim 1 or 2, further comprising split injection means for splitting and injecting fuel for one cycle injected by the first fuel injection valve into a plurality of times. A variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves for each valve;
Closing timing control means for delaying the closing timing of the second intake valve from the closing timing of the first intake valve;
The control apparatus for an internal combustion engine according to any one of claims 1 to 3, further comprising: Injection state changing means for executing fuel injection in a state where the fuel injection amount of the second fuel injection valve is set larger than the fuel injection amount of the first fuel injection valve;
Control switching means for operating the injection state change means instead of the injection state control means when raising the temperature of the first intake valve;
The control apparatus for an internal combustion engine according to any one of claims 1 to 4, further comprising:   The control switching means operates the injection state changing means when the air-fuel ratio of the internal combustion engine is in the vicinity of the stoichiometric air-fuel ratio, and when the air-fuel ratio is leaner than the stoichiometric air-fuel ratio, 6. The control apparatus for an internal combustion engine according to claim 5, wherein the control means is operated. A lift amount variable mechanism capable of variably setting a lift amount of the second intake valve;
Starting lift amount reducing means for reducing the lift amount of the second intake valve at the time of starting the internal combustion engine compared to the lift amount after starting;
The control device for an internal combustion engine according to any one of claims 1 to 6, further comprising: An ignition delay means for retarding the ignition timing at the start of the internal combustion engine;
The injection state control means, the control device for an internal combustion engine according to claim 1 or 4 comprising a structure for injecting respective fuel into a first half and a second half of the intake stroke by the at the start of the internal combustion engine first fuel injection valve. First and second exhaust ports provided at positions facing the first and second intake ports;
First and second exhaust valves provided in the first and second exhaust ports;
A variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves and the first and second exhaust valves for each valve;
Of the first and second intake valves and the first and second exhaust valves, only the first and second intake valves are opened in the first half of the intake stroke, and only the first exhaust valve is opened. A valve opening period control means that opens in the latter half of the intake stroke;
The control device for an internal combustion engine according to claim 1, comprising: First and second exhaust ports provided at positions facing the first and second intake ports;
First and second exhaust valves provided in the first and second exhaust ports;
A variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves and the first and second exhaust valves for each valve;
Valve opening period control means for opening the first and second intake valves over the first half and the second half of the intake stroke, and opening the first exhaust valve in the second half of the intake stroke;
3. The control apparatus for an internal combustion engine according to claim 1, wherein the injection state control means is configured to inject fuel in the first half and the second half of an intake stroke by the first fuel injection valve. A shut-off valve capable of closing the exhaust port except for a portion corresponding to the inner and outer circumferences of the first exhaust port, and having an exhaust partial shut-off valve that is driven to the closing side in the latter half of the intake stroke The control device for an internal combustion engine according to claim 9 or 10 . A variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves for each valve;
An opening timing control means for delaying the opening timing of the second intake valve from the opening timing of the first intake valve in the intake stroke;
The injection state control means is in a state where the first intake valve is opened and the second intake valve is opened after the first intake port is shut off by the intake adjustment valve. The control device for an internal combustion engine according to claim 1 or 2, wherein fuel is injected by the first fuel injection valve. An exhaust valve provided at the exhaust port;
A variable valve mechanism capable of independently changing the valve opening characteristics of the first and second intake valves and the exhaust valves for each valve;
Closing the exhaust valve before exhaust top dead center, opening the first intake valve between exhaust valve closing and exhaust top dead center, and And an opening / closing timing control means for opening the intake valve of 2 after exhaust top dead center,
The injection state control means is configured such that the first intake port is shut off by the intake adjustment valve and the first intake valve is opened until the exhaust top dead center arrives. The control device for an internal combustion engine according to claim 1 or 2, wherein fuel is injected by one fuel injection valve. The injection state control means includes
Injection means before and after top dead center for injecting fuel before and after exhaust top dead center by the first fuel injection valve;
If the engine speed is above the reference value, increase the fuel injection amount before exhaust top dead center over the fuel injection amount after exhaust top dead center, and if the engine speed is below the reference value, Injection amount control means for increasing the fuel injection amount after the dead center more than the fuel injection amount before the exhaust top dead center;
The control apparatus for an internal combustion engine according to claim 13 , comprising: The control device for an internal combustion engine according to claim 13 or 14 , further comprising self-ignition combustion control means for self-igniting the air-fuel mixture in the cylinder by compression. The first intake port and the first fuel injection valve are arranged on one side with respect to a reference straight line passing through the center of the combustion chamber in plan view, and the second intake port and the second fuel injection valve The valve is arranged on the other side with respect to the reference straight line,
The control apparatus for an internal combustion engine according to claim 1 or 2, wherein at least a fuel injection direction of the first fuel injection valve is inclined toward a center side of the combustion chamber. First and second exhaust ports provided at positions facing the first and second intake ports in a direction parallel to the reference straight line;
First and second exhaust valves provided in the first and second exhaust ports,
The fuel injection angle, which is an angle at which the fuel injected from the first fuel injection valve expands in plan view, is such that the center of the second intake valve and the center of the first exhaust valve are inside the fuel injection angle. The control device for an internal combustion engine according to claim 16, wherein the control device is set to an angle that falls within the range. The spray center line of the fuel injected from the first fuel injection valve according to claim than the center of the combustion chamber in plan view formed by arranging so as to pass through the position closer to the second fuel injection valve 16 Or the control device for an internal combustion engine according to 17 ; When the internal combustion engine is operated in a high rotation and high load state, the first and second fuel injection valves operate in place of the injection state control means and the intake adjustment valve is opened. The control device for an internal combustion engine according to any one of claims 1 to 18 , further comprising high-power control means for performing fuel injection by means of. The injection state control means drives the intake adjustment valve to the open side as the load of the internal combustion engine increases, and increases the fuel injection amount of the second fuel injection valve, so that the first and second fuels The control device for an internal combustion engine according to any one of claims 1 to 18 , wherein the control unit is configured to reduce a difference in fuel injection amount by an injection valve.

Images