JP2009162113A - Control device of internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To enhance a swirl ratio while suppressing the deterioration of a flow coefficient of an intake port, in a control device of an internal combustion engine. <P>SOLUTION: This control device has a first intake port (a main flow port) 60 having characteristics in which a swirl ratio is increased as an air amount passing through the inside is increased and an intake variable valve mechanism 40 making variable valve opening characteristics of an intake valves 64, 66. When a demand to increase the swirl ratio exists, both valve (all valve) early opening control for advancing opening times of all the intake valves 64, 66 arranged in the same cylinder is executed. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a control device for an internal combustion engine.

  Conventionally, for example, Patent Document 1 discloses an engine control device including an electromagnetically driven intake valve. This conventional control device includes valve timing correction amount calculation means for setting a phase difference in the intake air so that the strength of the swirl becomes a required value.

JP 2000-192820 A JP 2004-36566 A JP 2000-328968 A JP 2006-274777 A JP 2007-56701 A

  According to the conventional technique described above, a desired swirl strength can be obtained by providing a difference in flow rate of intake air between a plurality of intake ports. However, such a conventional technique has a problem that when the swirl strength is increased, the flow rate coefficient of the intake port decreases.

  The present invention has been made to solve the above-described problems, and an object of the present invention is to provide a control device for an internal combustion engine that can improve the swirl ratio while suppressing a decrease in the flow coefficient of the intake port.

A first invention is a control device for an internal combustion engine,
An intake port having at least one port having a characteristic that the swirl ratio increases as the amount of air passing through the interior increases;
An intake variable valve mechanism capable of changing a valve opening characteristic of an intake valve disposed in the intake port;
An intake valve control means for performing all-valve quick opening control for advancing the opening timing of all intake valves arranged in the same cylinder when there is a demand to increase the swirl ratio;
It is characterized by providing.

The second invention is the first invention, wherein
At least two intake valves are arranged in the cylinder,
The intake port includes a first intake port in which one of the at least two intake valves is disposed, and a second intake port in which the other of the at least two intake valves is disposed,
The intake variable valve mechanism is a mechanism capable of imparting different valve characteristics between the first intake port and the second intake port,
When the intake valve control means determines that the required swirl ratio is not reached even by the execution of the full-valve rapid opening control, the intake valve control unit is configured to perform intake air between the first intake port and the second intake port. It includes flow rate difference generation control means for performing flow rate difference generation control for controlling the intake variable valve mechanism so that a flow rate difference occurs in the intake air flowing through the port.

The third invention is the second invention, wherein
The flow rate difference generation control is a one-valve speeding up that closes one of the intake valves arranged in the first intake port and the intake valve arranged in the second intake port earlier than the other close timing. It is characterized by close control.

Moreover, 4th invention is 2nd invention.
The flow rate difference generation control is a one-valve delay that delays the opening timing of one of the intake valve arranged in the first intake port and the intake valve arranged in the second intake port compared to the opening timing of the other. It is characterized by opening control.

According to a fifth invention, in any one of the second to fourth inventions,
When the flow rate difference generation control means determines that the required swirl ratio is not reached even by the execution of the flow rate difference generation control, the flow rate difference generation control means controls the intake valve and the second intake port arranged in the first intake port. The intake valve is arranged so that the opening timing of one of the arranged intake valves is delayed compared to the opening timing of the other, and the closing timing of the one intake valve is advanced compared to the closing timing of the other intake valve. One-way slow opening and closing control for controlling the variable valve mechanism is performed.

The sixth invention is the fourth or fifth invention, wherein
Of the first intake port and the second intake port, the intake port on which the other intake valve is disposed is a port having higher swirl generation capability than the other intake port. And

The seventh invention is the sixth invention, wherein
The intake variable valve mechanism is a drive cam for driving the other intake valve disposed in the port having a high swirl generation capability, and the opening timing of the one intake valve is compared with the opening timing of the other intake valve. A one-valve quick-open drive cam used to drive the other intake valve independently of the one intake valve when the intake variable valve mechanism is controlled to be delayed.
The one-valve quick-open drive cam is fixed to the camshaft at a position on the advance side with respect to the drive cam used to drive the one intake valve.

The eighth invention is the second invention, wherein
The flow rate difference generation control is a one-valve speed-up method in which the opening timing of either one of the intake valve arranged in the first intake port and the intake valve arranged in the second intake port is advanced compared to the opening timing of the other. It is characterized by opening control.

The ninth invention is the eighth invention, wherein
When the flow rate difference generation control means determines that the required swirl ratio is not reached even by the execution of the first flow rate difference generation control, the intake valve and the second intake air that are disposed in the first intake port are determined. The opening timing of one of the intake valves arranged in the port is advanced compared to the opening timing of the other, and the closing timing of the one intake valve is delayed compared to the closing timing of the other intake valve. One-way valve early opening / closing control for controlling the intake variable valve mechanism is performed.

The tenth invention is the eighth or ninth invention, wherein
Of the first intake port and the second intake port, the intake port on which the one intake valve is disposed is a port having a higher swirl generation capability than the other intake port. And

The eleventh aspect of the invention is the tenth aspect of the invention,
The intake variable valve mechanism, as a cam for driving the one intake valve arranged in the port having a high swirl generation capability, compares the opening timing of the one intake valve with the opening timing of the other intake valve. A one-valve quick-open drive cam used to drive the one intake valve independently of the other intake valve when controlling the intake variable valve mechanism so as to accelerate
The one-valve quick-opening drive cam is fixed to the camshaft at a position on the advance side with respect to the drive cam used to drive the other intake valve.

The twelfth invention is the first invention, wherein
The intake valve control means performs the all-valve quick-open control when the operating range of the internal combustion engine is in a medium-high load range.

The thirteenth invention is the first or twelfth invention,
The internal combustion engine further includes an exhaust variable valve mechanism that can change a valve opening characteristic of the exhaust valve,
The intake valve control means opens the opening timing of all intake valves arranged in the cylinder when the operating range of the internal combustion engine is in the low load range, and opens when the operating range of the internal combustion engine is in the medium load range. Perform full valve slow-open control to control the value to the retard side of the timing,
The control apparatus for an internal combustion engine sets a closing timing of the exhaust valve when the operating region of the internal combustion engine is in a low load region, and is on a more retarded side than a closing timing when the operating region of the internal combustion engine is in a medium load region. It further comprises exhaust valve control means for performing exhaust slow closing control to be controlled to a value.

The fourteenth invention is the thirteenth invention, in which
The intake valve control means determines that the discharge of unburned HC cannot be sufficiently suppressed even when the all-valve slow opening control and the exhaust slow closing control are executed when the operating range of the internal combustion engine is in a low load range. In this case, the closing timing of all the intake valves is controlled to the timing near the intake bottom dead center.

In addition, a fifteenth aspect of the invention is any one of the first, twelfth to fourteenth aspects of the invention,
The intake valve control means performs the all-valve quick-open control when the operating range of the internal combustion engine is in a high rotation and high load range,
The intake valve control means further includes any one of an intake valve disposed in the first intake port and an intake valve disposed in the second intake port when the operating region of the internal combustion engine is in a high rotation / high load region. One-valve early closing control is performed in which the closing timing of one of them is advanced compared to the closing timing of the other.

The sixteenth aspect of the invention is the fifteenth aspect of the invention,
The control apparatus for an internal combustion engine further includes intake pressure acquisition means for acquiring pressure in the intake passage of the internal combustion engine, and exhaust pressure acquisition means for acquiring pressure in the exhaust passage of the internal combustion engine,
The intake valve control means decreases the advance amount of the intake valve opening timing by the all-valve early opening control as the difference between the exhaust pressure and the intake pressure increases.

According to a seventeenth aspect, in the first aspect,
The intake valve control means includes a cold-time intake valve control means for reducing the advance amount of the opening timing of the intake valve by the full valve early opening control as the cooling water temperature of the internal combustion engine becomes lower. .

According to an eighteenth aspect, in the first aspect,
The internal combustion engine further includes an exhaust variable valve mechanism that can change a valve opening characteristic of the exhaust valve,
The control device for an internal combustion engine further includes cold exhaust valve control means for increasing a retard amount of the closing timing of the exhaust valve as the cooling water temperature of the internal combustion engine becomes lower.

The nineteenth aspect of the invention is the eighteenth aspect of the invention,
The exhaust port in which the exhaust valve is disposed has a direction opposite to an intake swirl flow generated by at least one of the first intake port and the second intake port when exhaust gas flows backward from the exhaust port into the cylinder. It is a port that generates an exhaust swirl flow.

The twentieth invention is the eighteenth invention, in the eighteenth invention,
The exhaust port in which the exhaust valve is disposed is a swirl in which at least one of the first intake port and the second intake port generates the exhaust gas that flows backward when the exhaust gas flows backward from the exhaust port into the cylinder. It is a straight port arranged in such a direction as to collide with a flow.

The twenty-first invention is the first invention, wherein
The intake valve control means includes an acceleration intake valve control means for advancing the opening timing of the at least one intake valve in order to increase a swirl ratio during acceleration.

  According to 1st invention, when there exists a request | requirement which raises a swirl ratio, a swirl ratio can be raised, without causing the fuel consumption deterioration by the fall of a flow coefficient (increase in pump loss).

  According to the second aspect of the invention, when the swirl ratio is increased, execution of the all-valve rapid opening control has priority over the flow rate difference generation control that causes a flow rate difference in the intake air between the first intake port and the second intake port. Will come to be. As a result, by providing a flow rate difference in the intake air between the first and second intake ports, the swirl ratio can be sufficiently increased, but an opportunity to execute flow rate difference generation control accompanied by a decrease in the flow coefficient (increase in pump loss). Can be kept to the minimum necessary. In addition, since the full valve quick opening control is utilized to the maximum extent, the amount of use of the flow rate difference generation control when increasing the swirl ratio can be reduced. For this reason, it becomes possible to acquire an appropriate swirl ratio according to the degree of demand while suppressing deterioration in fuel consumption due to a decrease in the flow coefficient.

  According to the third aspect, by adjusting the valve opening characteristics of the intake valve, it is possible to satisfactorily produce a flow rate difference in the intake air flowing through the first and second intake ports.

  According to the fourth invention, by adjusting the valve opening characteristics of the intake valve, it is possible to satisfactorily produce a flow rate difference in the intake air flowing through the first and second intake ports.

  According to the fifth aspect, when the swirl ratio is increased, the execution of the flow rate difference generation control is prioritized over the one-valve slow opening / closing control. As a result, it is possible to acquire an appropriate swirl ratio according to the degree of demand while suppressing deterioration in fuel consumption due to a decrease in the flow coefficient.

  According to the sixth aspect of the invention, the other intake valve arranged in the intake port having the higher swirl generation capability is opened earlier than the other intake valve. For this reason, in the initial stage of the intake stroke, the flow of intake air flowing through the intake port having the higher swirl generation capability can be increased, and the swirl ratio can be increased efficiently.

  According to the seventh aspect of the invention, the setting of the one-valve quick-open drive cam makes it possible to increase the flow of intake air flowing through the intake port with the higher swirl generation capability at the initial stage of the intake stroke.

  According to the eighth aspect, by adjusting the valve opening characteristics of the intake valve, it is possible to satisfactorily produce a flow rate difference in the intake air flowing through the first and second intake ports.

  According to the ninth aspect, when the swirl ratio is increased, the execution of the flow rate difference generation control is prioritized over the one-valve quick opening / closing control. As a result, it is possible to acquire an appropriate swirl ratio according to the degree of demand while suppressing deterioration in fuel consumption due to a decrease in the flow coefficient.

  According to the tenth aspect of the invention, one intake valve disposed in the intake port having the higher swirl generation capability is opened earlier than the other intake valve. For this reason, in the initial stage of the intake stroke, the flow of intake air flowing through the intake port having the higher swirl generation capability can be increased, and the swirl ratio can be increased efficiently.

  According to the eleventh aspect of the invention, the setting of the one-valve quick-open drive cam makes it possible to increase the flow of the intake air flowing through the intake port having the higher swirl generation capability at the initial stage of the intake stroke.

  According to the twelfth aspect of the present invention, the swirl ratio can be increased without causing deterioration in fuel consumption due to a decrease in the flow coefficient (increase in pump loss) in a medium to high load region where a high swirl ratio is required.

  According to the thirteenth aspect of the present invention, in the low load region where the swirl ratio is required to be lower than that in the middle and high load region, the swirl is performed by retarding the opening timing of the intake valve and retarding the closing timing of the exhaust valve. Can be suppressed.

  According to the fourteenth aspect of the invention, when the swirl ratio is reduced in the low load region, the execution of the exhaust exhaust gas closing control has priority over the control of the closing timing of the intake valve to the timing near the intake bottom dead center. It becomes like this. As a result, when the intake valve has a characteristic that the operating angle of the intake valve decreases by advancing the closing timing of the intake valve, the unburned HC is discharged while suppressing the deterioration of the fuel consumption caused by the decrease of the operating angle of the intake valve. It becomes possible to suppress appropriately according to a request | requirement degree.

  According to the fifteenth aspect of the present invention, in the high-rotation and high-load region where a high swirl ratio is required, by executing both the single valve rapid opening control and the single valve rapid closing control, an appropriate swirl ratio corresponding to the required degree can be obtained. It becomes possible to acquire enough.

  According to the sixteenth aspect of the present invention, in a situation where there is a concern that the amount of internal EGR gas increases due to the return of exhaust gas from the exhaust side to the intake side due to advancement of the opening timing of the intake valve. An increase in the amount of internal EGR gas due to slow blow-back can be suppressed. For this reason, as a result of performing control to increase the swirl ratio for the purpose of suppressing smoke emission, it is possible to avoid the smoke emission amount from increasing as the internal EGR gas amount increases.

  According to the seventeenth aspect of the present invention, fuel consumption is improved by appropriately suppressing the THC discharge amount under a low cooling water temperature condition, and by increasing the opening timing of the intake valve as the cooling water temperature rises. Can be planned. As described above, even in the warm-up process, by appropriately controlling the opening timing of the intake valve, it is possible to reduce both the THC emission amount and the fuel consumption, which are in a trade-off relationship, to the maximum.

  According to the eighteenth aspect, the swirl can be appropriately suppressed according to the cooling water temperature by adjusting the retard amount of the exhaust valve closing timing. Thereby, the discharge of THC can be suitably suppressed when cold. Further, the swirl ratio can be controlled only by adjusting the valve opening characteristics of the exhaust valve.

  According to the nineteenth aspect of the invention, when the swirl is suppressed by adjusting the retard amount of the exhaust valve closing timing, the generation of the intake swirl flow can be more effectively suppressed, and the THC can be sufficiently discharged. It becomes possible to suppress.

  According to the twentieth aspect, when the swirl is suppressed by adjusting the retard amount of the exhaust valve closing timing, the generation of the intake swirl flow can be suppressed more effectively, and the THC can be sufficiently discharged. It becomes possible to suppress. Further, since the exhaust port is configured as a straight port, the intake swirl flow can be adjusted without deteriorating the exhaust efficiency as compared with a case where a helical port is used, for example.

  According to the twenty-first aspect, it is possible to increase the swirl ratio without deteriorating the flow coefficient. That is, it is possible to improve both the swirl ratio and the intake air amount. For this reason, it is possible to increase the fuel injection amount by improving the swirl ratio without the contradiction of reducing the amount of fresh air sucked into the cylinder. This makes it possible to improve the response of the internal combustion engine during acceleration.

Embodiment 1 FIG.
[Description of system configuration]
FIG. 1 is a diagram for explaining a system configuration according to the first embodiment of the present invention. The system shown in FIG. 1 includes an internal combustion engine 10. The internal combustion engine 10 is a four-cycle diesel engine (compression ignition type internal combustion engine). Each cylinder of the internal combustion engine 10 is provided with an injector 12 that directly injects fuel into the cylinder. The injectors 12 of each cylinder are connected to a common common rail 14. Fuel in a fuel tank (not shown) is pressurized to a predetermined fuel pressure by a supply pump 16, stored in the common rail 14, and supplied from the common rail 14 to each injector 12.

  An exhaust passage 18 of the internal combustion engine 10 is branched by an exhaust manifold 20 and connected to exhaust ports 68 and 70 (see FIG. 2) of each cylinder. The internal combustion engine 10 of the present embodiment includes a turbocharger 22. The exhaust passage 18 is connected to the exhaust turbine of the turbocharger 22. Further, an exhaust gas purification device 24 for purifying exhaust gas is provided on the exhaust passage 18 downstream of the turbocharger 22. The exhaust manifold 20 is provided with an exhaust pressure sensor 58 that detects the pressure in the exhaust manifold 20.

  An air cleaner 28 is provided near the inlet of the intake passage 26 of the internal combustion engine 10. The air sucked through the air cleaner 28 is compressed by the intake compressor of the turbocharger 22 and then cooled by the intercooler 30. The intake air that has passed through the intercooler 30 is distributed by the intake manifold 32 to the intake ports 60 and 62 (see FIG. 2) of each cylinder.

  An intake throttle valve 34 is installed between the intercooler 30 and the intake manifold 32. An air flow meter 36 for detecting the intake air amount is installed in the vicinity of the downstream side of the air cleaner 28. An intake pressure sensor 38 that detects the pressure in the intake passage 26 is installed on the downstream side of the intake throttle valve 34.

  Further, the system shown in FIG. 1 includes an intake variable valve mechanism 40 that can change the valve opening characteristics of the intake valves 64 and 66 (see FIG. 2) of each cylinder, and exhaust valves 72 and 74 (see FIG. 2). An exhaust variable valve mechanism 42 that can change the valve opening characteristics is provided.

More specifically, the intake variable valve mechanism 40 includes a function (two-valve variable function) capable of continuously changing two intake valves 64 and 66 arranged per cylinder with the same valve opening characteristics, and one intake valve. A function (single valve variable function) capable of continuously changing the valve opening characteristic of the other intake valve 66 while fixing the lift amount (and operating angle) of the valve 64 to a predetermined high lift amount (and large operating angle); It shall be equipped with. The intake variable valve mechanism 40 having these functions can be realized, for example, by adopting the configuration of the variable valve mechanism detailed in International Publication No. WO 2006/132059. The intake variable valve mechanism 40 is also provided with a function (phase variable function) that can change the opening and closing timing of the intake valves 64 and 66 while keeping the lift amount and the operating angle constant. Such a phase variable function can be realized by, for example, providing a VVT mechanism (not shown) for controlling the opening / closing timing of the intake valves 64 and 66.
According to the intake variable valve mechanism 40 having such a configuration, the opening timing of the intake valves 64 and 66 can be controlled so as to be a value on the advance side of the intake top dead center by the phase variable function. It is also possible to perform control so that the closing timing of the intake valve 66 is earlier than the closing timing of the intake valve 64 by the single valve variable function (single valve early closing control).

  The exhaust variable valve mechanism 42 is assumed to have a function (phase variable function) that can change the opening / closing timing of the exhaust valves 72 and 74 while keeping the lift amount and the working angle constant. Such a function can be realized, for example, by providing a VVT mechanism (not shown) for controlling the opening / closing timing of the exhaust valves 72 and 74.

  Further, the system of the present embodiment includes an ECU (Electronic Control Unit) 50. In addition to the various sensors described above, the ECU 50 includes a crank angle sensor 52 for detecting the engine speed, an accelerator opening sensor 54 for detecting the accelerator opening, and a coolant temperature of the internal combustion engine 10. The water temperature sensor 56 is connected to the above-described various actuators. The ECU 50 controls the operating state of the internal combustion engine 10 by driving each actuator according to a predetermined program based on those sensor signals and information.

[Configuration around intake and exhaust ports]
FIG. 2 is a view for explaining a specific configuration around the intake and exhaust ports in the internal combustion engine 10 shown in FIG.
As shown in FIG. 2, the internal combustion engine 10 includes two intake ports 60 and 62. A first intake valve 64 and a second intake valve 66 are disposed in the first intake port 60 and the second intake port 62, respectively. The internal combustion engine 10 includes two exhaust ports 68 and 70. A first exhaust valve 72 and a second exhaust valve 74 are disposed in the first exhaust port 68 and the second exhaust port 70, respectively. In addition, the arrow attached | subjected to each port 60 grade | etc., In FIG. 2 has shown the direction of the flow of the said gas when gas is introduce | transduced in a cylinder from each port 60 grade | etc.,.

  The first intake port 60 is configured as a port having a higher swirl generation capability than the second intake port 62, in other words, a port mainly responsible for swirl generation (mainstream port). More specifically, such a port with high swirl generation capability can be realized, for example, by configuring the first intake port 60 as a tangential port. According to the first intake port 60 configured as such a tangential port, as shown in FIG. 2, since the first intake port 60 is along the tangential direction of the cylinder, it is possible to effectively swirl the cylinder. In addition, a port characteristic can be obtained in which the swirl ratio increases as the amount of air passing through the interior increases.

  Further, as shown in FIG. 2, the first exhaust port 68 has an intake swirl due to a gas stream flowing backward from the first exhaust port 68 into the cylinder through the first intake port 60. It is configured to collide with streamlines. According to such a configuration of the first exhaust port 68, the exhaust valve 72, 74 is closed more slowly than the intake top dead center, so that the exhaust gas flows backward from the first exhaust port 68 into the cylinder. Under the conditions that occur, the intake swirl flow can be weakened by the exhaust gas that flows back into the cylinder.

[Advantages of swirl improvement by quick opening control of both valves (all valves)]
FIG. 3 is a diagram for explaining the advantage of swirl improvement by the both-valve rapid opening control.
More specifically, FIG. 3 (A) shows the relationship when the swirl improvement technique using the one-valve early closing control is used. When the swirl is increased by providing a flow rate difference in the intake air between the two intake ports by the single valve early closing control, the intake flowing through the second intake port (subordinate port) 62 on the intake valve 66 side that is quickly closed. Since the amount of air is limited, as shown in FIG. 3A, the flow rate coefficients of the intake ports 60 and 62 and the swirl ratio basically have a trade-off relationship. That is, basically, when the flow coefficient is increased, the swirl ratio becomes small (the swirl becomes weak), and conversely, when the swirl ratio is increased, the flow coefficient becomes small. Such a relationship is the same in the control using the swirl control valve (SCV).

On the other hand, FIG. 3 (B) shows the relationship when the swirl improvement method using the double valve quick opening control is used. The data shown in FIG. 3 (B) is based on conditions where the operating angles of the intake valves 64 and 66 are constant.
If both the opening timings of the two intake valves 64 and 66 are controlled to a value more advanced than the intake top dead center by such double valve early opening control, the initial stage of the intake stroke (after the top dead center) In the range of 0 to 90 ° CA), that is, when the amount of gas sucked into the cylinder increases, the lift amount of the intake valves 64 and 66 is increased. As a result, the intake air is vigorously drawn into the cylinder, so that the swirl can be increased. According to such double valve quick opening control, as shown in FIG. 3B, when the advance amount of the intake valves 64 and 66 is increased, the flow rate coefficient is deteriorated (that is, the pump loss is reduced). The swirl ratio can be increased without incurring an increase.

  The first intake port 60 is configured as a tangential port so that the swirl ratio can be effectively increased when the double valve quick opening control is executed. In addition, the first intake port 60 is the main intake port. The port 60 may be configured as a helical port formed in a spiral shape, for example. Further, in order to effectively increase the swirl ratio during the execution of the double valve quick opening control, instead of the above configuration, for example, the first intake port 60 which is a mainstream port is configured as a tangential port. The second intake port 62, which is the other dependent port, may be configured as a helical port.

[Swirl control according to Embodiment 1 mainly using double valve quick opening control]
By the way, when the operating region of the internal combustion engine 10 is in the middle and high load region, a high swirl ratio is required to suppress smoke discharge. On the other hand, when the operating region of the internal combustion engine 10 is in the low load region, if the swirl ratio is sufficiently increased, the spray in the cylinder is excessively diffused, and a large amount of unburned HC (THC) is discharged. End up.

  Therefore, in the present embodiment, an appropriate swirl ratio corresponding to each operation region of the internal combustion engine 10 is obtained while suppressing a decrease in the flow coefficient of the intake ports 60 and 62 (that is, a deterioration in fuel consumption accompanying an increase in pump loss) as much as possible. In order to achieve this, as shown in FIG. 4 below, the opening / closing timing of the intake / exhaust valve 64 and the like is adjusted in accordance with the operating region of the internal combustion engine 10.

FIG. 4 is a diagram for explaining each control method of the intake / exhaust valve 64 and the like used in the first embodiment of the present invention, and FIG. 5 uses each method of control A to F shown in FIG. FIG. 3 is a diagram showing an operating region in relation to the torque of the internal combustion engine 10 and the engine speed.
Each waveform of the controls A to F shown in FIG. 4 indicates a valve lift curve of the intake / exhaust valve 64 or the like. Also, the valve lift curve shown in FIG. 4D is a base valve lift curve used in a predetermined medium load region as shown in FIG. More specifically, in the lift curve of the base shown in FIG. 4D, the intake valves 64 and 66 are later than the intake bottom dead center (BDC) after opening at a timing near the intake top dead center (TDC). The exhaust valves 72 and 74 are set to close at a timing near the intake top dead center (TDC) after being opened at a predetermined timing later in the expansion stroke. .

  In the middle and high load region of the internal combustion engine 10, basically, a higher swirl ratio is required as the load increases. Therefore, in the present embodiment, as shown in FIG. 5, basically, both the intake valves 64 and 66 are opened in a medium and high load region where a higher swirl ratio is required than in a medium load region where the base control D is used. Both valve early opening control is performed to advance the timing.

  More specifically, when a swirl ratio higher than the usage range of the base control D is required, the control C that performs only the two-valve quick-opening control with respect to the base control D (see FIG. 4C). ). In addition, when the control C is executed, the advance amount of the intake valves 64 and 66 is increased as the required swirl ratio becomes higher.

  In a high load range where the required swirl ratio cannot be ensured only by executing the double valve rapid opening control as described above, basically, the single valve rapid opening control is performed in addition to the dual valve rapid opening control. The control A to be performed (see FIG. 4A) is executed. When this control A is executed, the closing timing of the intake valve 66 on the early closing side is controlled to a more advanced value as the required swirl ratio becomes higher.

  Further, when control A is executed, in the region where the difference between the exhaust pressure (exhaust manifold pressure) and the supercharging pressure (intake manifold pressure) is large (existing in the high rotation and high load region), the double valve quick opening control is executed. And the control B (see FIG. 4B) for performing only the one-valve early closing control is executed.

  Further, in the present embodiment, as shown in FIG. 5, a large amount of THC may be discharged when the swirl ratio is increased in the region on the lower load side than the region where the base control D is used. Since there is a concern, basically, the control E for performing the exhaust slow closing control for prohibiting the execution of the double valve early opening control and delaying the closing timing of the exhaust valves 72 and 74 from the intake top dead center (FIG. 4E). See). In addition, in the case of an extremely low load region where it can be determined that THC cannot be sufficiently reduced in the low load region only by executing the exhaust exhaust valve closing control, the closing timing of both intake valves 64 and 66 in addition to the exhaust exhaust valve closing control. The control F (see FIG. 4F) is performed to perform the intake early closing control for controlling the intake to close at the timing near the intake bottom dead center.

  As described above, according to the control of the present embodiment, the swirl ratio becomes smaller as the alphabet attached to the lift curve becomes younger than the base lift curve as compared with the case where the base lift curve is used. It becomes higher, THC deteriorates, and smoke emission becomes improved.

Next, with reference to FIG. 6, the specific process in Embodiment 1 of this invention is demonstrated. FIG. 6 is a flowchart of a routine executed by the ECU 50 in the first embodiment to realize the above function.
In the routine shown in FIG. 6, it is first determined whether or not the current load region of the internal combustion engine 10 is a base load region (step 100). As a result, when it is determined that the current load region is the base load region, the control D employing the base valve lift curve shown in FIG. 4D is selected (step 102).

  On the other hand, if it is determined in step 100 that the current load region is not the base load region, it is determined whether or not the current load region is higher than the base load region (step 104). As a result, when it is determined that the current load region is higher than the base load region, it is determined that there is a request to increase the swirl ratio, and first, control C is selected, and control D is used as a reference. Both-valve rapid opening control is executed to open the intake valves 64 and 66 earlier than the intake top dead center (step 106). More specifically, switching from the control D to the control C is realized by a phase variable function provided in the intake variable valve mechanism 40.

  Next, it is determined whether or not the required swirl ratio is achieved by the swirl generated only by the execution of the double valve quick opening control (step 108). More specifically, whether or not the currently generated swirl satisfies the required swirl ratio corresponding to the operating region of the internal combustion engine 10 is determined from the internal combustion engine 10 by using, for example, a smoke sensor (not shown). This can be determined by obtaining the amount of smoke discharged. That is, for example, when the swirl generated only by executing the double-valve quick-opening control is not sufficiently reduced to the target level of smoke discharge, it is generated only by executing the double-valve quick-opening control. Depending on the swirl, it can be determined that the required swirl ratio cannot be achieved.

  If it is determined in step 108 that the required swirl ratio is achieved by execution of control C, control C remains selected, while the required swirl ratio is achieved only by execution of control C. If it is determined that there is not, then it is determined whether or not the pressure difference between the exhaust pressure and the supercharging pressure is greater than a predetermined determination value (step 110).

  As a result, when it is determined that the pressure difference> the predetermined value is not satisfied, the control A is selected instead of the control C in order to execute the one-valve early closing control in addition to the both-valve rapid opening control. (Step 112). More specifically, switching from the control C to the control A is performed by executing the one-valve early closing control in a state where the opening timing of the intake valves 64 and 66 is advanced from the intake top dead center. Realized.

  On the other hand, when it is determined that the pressure difference> the predetermined value is established, the control B is selected in place of the control C so as to prohibit the execution of the double valve quick opening control and execute only the single valve quick closing control. (Step 114). More specifically, the switching from the control C to the control B is performed by executing the one-valve early closing control while returning the opening timing of the intake valves 64 and 66 to the timing near the intake top dead center by the phase variable function. Realized.

  On the other hand, if it is determined in step 104 that the current load area is not higher than the base load area, that is, the current load area is lower than the base load area, the control E selects Then, exhaust exhaust gas closing control is executed (step 116). In other words, in this case, the both-valve quick-open control is not executed, and the opening timing of both intake valves 64 and 66 is retarded as compared to the case where it is a region higher than the base load region. More specifically, the exhaust exhaust closing control in the control E is realized by the phase variable function of the exhaust variable valve mechanism 42.

  Next, it is determined whether or not the THC emission amount is sufficiently suppressed by the execution of the control E (step 118). More specifically, whether or not the current THC emission amount is sufficiently suppressed is, for example, temperature information of the exhaust purification device 24 (catalyst) using a temperature sensor (not shown), an in-cylinder pressure sensor, or the like. This can be determined by acquiring the current THC emission amount based on the combustion state information of the internal combustion engine 10 using (not shown).

  If it is determined in step 118 that the THC emission amount is sufficiently suppressed to the target level by execution of control E, control E remains selected, while only execution of control E is performed. If it is determined that the THC emission amount is not sufficiently suppressed, the control F is executed instead of the control E in order to execute the early intake closing control in addition to the exhaust late closing control (step 120). . More specifically, the intake early closing control in the control F is realized by the one-valve variable function of the intake variable valve mechanism 40.

  According to the routine shown in FIG. 6 described above, when a high swirl ratio is requested due to a request to shift to a region higher in load than the base load region, first, the double valve early opening control is performed. It will be executed preferentially. According to the both-valve quick-open control, as described above, the swirl ratio can be increased without causing deterioration in fuel consumption due to a decrease in flow coefficient (increase in pump loss). That is, it is possible to suppress the smoke emission amount without causing deterioration of fuel consumption.

  Then, according to the above routine, when it can be determined that the smoke discharge amount cannot be sufficiently suppressed only by the improvement in the swirl ratio by the execution of the double valve quick opening control, basically, in addition to the double valve quick opening control, Early valve closing control is executed. By performing the control that gives priority to the two-valve quick opening control in this way, the swirl ratio can be sufficiently increased by providing a flow rate difference in the intake air between the two intake ports 60, 62. This makes it possible to keep the opportunity for executing the one-valve early closing control accompanied by a decrease in the above (increase in pump loss) to a necessary minimum. In addition, since the double-valve rapid opening control is utilized to the maximum extent, it is possible to reduce the usage of the single-valve rapid closing control when increasing the swirl ratio. For this reason, it becomes possible to acquire an appropriate swirl ratio according to the degree of demand while suppressing deterioration in fuel consumption due to a decrease in the flow coefficient.

  Further, according to the above routine, in an area where a high swirl ratio is required, such as in a medium / high load area, the exhaust pressure and the supercharging are basically basically used together with the both-valve rapid opening control and the one-valve rapid closing control. When the difference from the pressure is larger than a predetermined value, only the one-valve early closing control is executed. Under a situation where the pressure difference between the exhaust pressure and the supercharging pressure is large, if the opening timing of the intake valves 64 and 66 is advanced from the intake top dead center by the double valve early opening control, the exhaust ports 68 and 70 are The amount of internal EGR gas increases as the exhaust gas blows back toward the intake ports 60 and 62. When the amount of internal EGR gas increases, the amount of fresh air sucked into the cylinder decreases, and the smoke discharge amount increases. According to the above routine, under the circumstances where the pressure difference is large, the execution of the two-valve rapid opening control is prohibited. Therefore, as a result of performing the control to increase the swirl ratio in order to suppress the discharge of smoke, It can be avoided that the discharge amount increases instead.

  Further, according to the above routine, when a request to shift to a lower load region than the base load region is issued, first, the exhaust late closing control is preferentially performed in a state in which the double valve early opening control is prohibited. Will be executed. According to the exhaust slow closing control, in the intake stroke, exhaust gas can be introduced into the cylinder from the exhaust ports 68, 70 side, and the gas flow introduced from the exhaust ports 68, 70 side and the intake ports 60, 62 side The gas flow from can be made to interfere in the cylinder. Thereby, it is possible to suppress the swirl while reducing the THC emission amount by increasing the internal EGR gas amount.

  Then, according to the above routine, in the extremely low load region where it is determined that the THC emission amount cannot be sufficiently suppressed only by executing the exhaust exhaust closing control, the intake early closing control is performed in addition to the exhaust exhaust closing control. Will be executed. If the intake valves 64 and 66 are closed at a timing in the vicinity of the intake bottom dead center by the intake early closing control, the THC exhaust amount can be satisfactorily reduced by the increase in the compression end temperature caused by the increase in the actual compression ratio. However, on the other hand, there is also one aspect that the fuel consumption deteriorates due to an increase in pump loss due to a decrease in the operating angle of the intake valves 64 and 66. For this reason, when the reduction of the THC emission amount is required in the low load region, the control that preferentially executes the exhaust late closing control among the exhaust late closing control and the intake early closing control as in the routine processing described above. By doing so, it is possible to minimize the opportunity for executing the intake early closing control, which is effective for reducing the THC emission amount but has one aspect of deterioration of fuel consumption. In addition, since exhaust exhaust closing control is used to the maximum extent, it is possible to reduce the amount of use of intake early closing control when suppressing THC emission. For this reason, it becomes possible to reduce THC satisfactorily while suppressing deterioration in fuel consumption.

  By the way, in the first embodiment described above, the both-valve quick-opening control is performed by utilizing the phase variable function of the intake variable valve mechanism 40. However, the present invention is not limited to this method, and a method of advancing only the opening timings of all the intake valves arranged in the same cylinder may be used. Similarly, the exhaust slow closing control is not limited to the phase variable function of the exhaust variable valve mechanism 42, and a method of delaying only the closing timing of the exhaust valve may be used.

  In the first embodiment described above, when the current load region of the internal combustion engine 10 is higher than the predetermined base load region, it is determined that there is a request to increase the swirl, and the both-valve quick opening control is performed. ing. However, the intake valve control means in the present invention is not limited to such a method, and for example, the advance amount of the opening timing of the intake valve in consideration of the swirl request is determined based on the operating range of the internal combustion engine 10 (torque, engine speed, etc.). In such a case, a predetermined map may be provided so that a swirl request is not directly or indirectly acquired on an actual machine.

In the first embodiment described above, the first intake port 60 corresponds to the “intake port” in the first invention. Further, the “intake valve control means” in the first aspect of the present invention is realized by the ECU 50 executing the processing of steps 104 and 106 described above.
Further, the “flow rate difference generation control means” in the first aspect of the present invention is realized by the ECU 50 executing the processing of steps 108 to 114 described above.
Further, the “exhaust valve control means” according to the thirteenth aspect of the present invention is realized by the ECU 50 executing the processing of steps 104 and 116-120.
The intake pressure sensor 38 and the exhaust pressure sensor 58 correspond to the “intake pressure acquisition means” and the “exhaust pressure acquisition means” in the sixteenth aspect of the invention, respectively.

Embodiment 2. FIG.
Next, a second embodiment of the present invention will be described with reference to FIGS.
The system of the present embodiment uses the hardware configuration shown in FIG. 1 (the intake variable valve mechanism uses an intake variable valve mechanism 80 described later), and the ECU 50 replaces the routine shown in FIG. This can be realized by executing the routine shown in FIG.

[Configuration of intake variable valve mechanism]
FIG. 7 is a perspective view for explaining a specific configuration of intake variable valve mechanism 80 according to Embodiment 2 of the present invention.
The basic configuration of the intake variable valve mechanism 80 of the present embodiment is basically the same as the variable valve mechanism detailed in International Publication No. WO 2006/132059 of the international application, except for the points described below. It is the same.
That is, the difference from the variable valve mechanism described in the above publication is that the second drive fixed at a position shifted to the advance side by a predetermined angle α with respect to the first drive cam 82 used during the variable valve control. The cam 84 (see FIG. 8 described later) drives the first intake valve 64 disposed in the first intake port 60, which is the intake port (mainstream port) having a higher swirl generation capability, during the one-valve early closing control. It is in the point of doing so. In addition, about the part same as the variable valve mechanism described in the said gazette, only the outline will be shown here and the detailed description shall be abbreviate | omitted or simplified.

  As shown in FIG. 7, the cam shaft 86 of the intake variable valve mechanism 80 is provided with two drive cams 82 and 84 per cylinder. Two intake valves 64 and 66 are arranged symmetrically about one drive cam 82 as a center. A first variable mechanism 88 is provided between the first drive cam 82 and the first intake valve 64 to link the lift movement of the first intake valve 64 with the rotational movement of the first drive cam 82. In addition, a second variable mechanism 90 is provided between the first drive cam 82 and the second intake valve 66 to link the lift movement of the second intake valve 66 with the rotational movement of the first drive cam 82. In addition, a fixed valve mechanism 92 is provided between the other drive cam 84 and the first intake valve 64 to link the lift movement of the first intake valve 64 with the rotational movement of the second drive cam 84. .

  As shown in FIG. 7, in the variable mechanisms 88 and 90, the rocker arm 94 is supported by the intake valves 64 and 66. The variable mechanisms 88 and 90 are respectively interposed between the first drive cam 82 and the rocker arm 94, and continuously change the interlocking state between the rotational motion of the first drive cam 82 and the rocking motion of the rocker arm 94. It is like that.

  These variable mechanisms 88 and 90 include a control shaft 96, a control arm 98, a link arm 100, a swing cam arm (first swing cam arm 102, second swing cam arm 104), and intermediate rollers (cam roller 106, arm roller 108). ) As a main constituent member.

  As shown in FIG. 7, the fixed valve mechanism 92 includes a large lift arm 110 driven by the second drive cam 84. The fixed valve mechanism 92 includes an arm coupling mechanism (not shown) that couples and releases the large lift arm 110 and the first swing cam arm 102 by inserting and removing a pin by hydraulic pressure. Here, the first swing cam arm 102 and the large lift arm 110 are coupled when the posture of the first swing cam arm 102 can sufficiently increase the operating angle and the lift amount of the first intake valve 64. It is supposed to be done.

  According to the intake variable valve mechanism 80 configured as described above, both are driven by the first drive cam 82 in a state where the coupling between the first swing cam arm 102 and the large lift arm 110 is released. The operating angle and lift amount of both intake valves 64 and 66 can be continuously varied by adjusting the rotational position of the control shaft 96 (both valve variable function).

  Further, according to the intake variable valve mechanism 80, the first intake valve 64 driven by the second drive cam 84 is in a state where the first swing cam arm 102 and the large lift arm 110 are coupled. The operating angle and lift amount of the other intake valve 66 driven by the first drive cam 82 are adjusted to adjust the rotational position of the control shaft 96 while fixing the operating angle and lift amount to a predetermined large operating angle and high lift amount. Can be continuously changed (single valve variable function).

  The intake variable valve mechanism 80 is also provided with a VVT mechanism (not shown) for adjusting the phase of the camshaft 86 with respect to the crankshaft at one end of the camshaft 86. A function (phase variable function) that can change the opening / closing timing without changing the lift amount and the operating angle can also be realized.

  FIG. 8 is a view of the intake variable valve mechanism 80 as viewed from the axial direction of the control shaft 96 on the arrow A side in FIG. More specifically, FIG. 8 shows that the pressing force of the first drive cam 82 acts on the swing cam arms 102 and 104 when the base circle portion of the first drive cam 82 is in contact with the cam roller 106. It is a figure when not doing.

  As shown in FIG. 8, the second drive cam 84 is fixed to the cam shaft 86 at a position shifted from the first drive cam 82 by a predetermined angle α toward the advance side. According to such setting of the second drive cam 84, the first mainstream port 60 having a high swirl generation capability is used when the one-valve variable function for variably controlling the operating angle and lift amount of only the second intake valve 66 is used. Therefore, the opening timing of the first intake valve 64 on the fixed side arranged in the position is advanced earlier than the opening timing of the second intake valve 66 on the variable side.

  For this reason, the one-valve variable function in this case can be arranged as follows. That is, when focusing on the first intake valve 64 in a situation where such a one-valve variable function is used, the first intake valve 64 is opened more quickly than the second intake valve 66, and the second As compared with the intake valve 66, it is closed late. Therefore, in this specification, the control using the one-valve variable function in this case may be referred to as “one-valve quick opening / closing control”. On the other hand, when focusing on the second intake valve 66, the second intake valve 66 is opened more slowly than the first intake valve 64 and closed earlier than the first intake valve 64. . Therefore, in this specification, the control using the one-valve variable function in this case may be referred to as “one-valve slow opening / closing control” in other words.

[Control of Intake Valve in Embodiment 2]
FIG. 9 is a diagram for explaining the swirl improvement effect by the one-valve rapid opening / closing control realized by the intake variable valve mechanism 80 in the second embodiment of the present invention.
In FIG. 9, first, only the one-valve early closing control (white circle mark) is compared with the both-valve early opening control + one-valve early closing control (white triangle mark). When both valve early opening control is executed in addition to single valve early closing control, the intake air amount in the initial stage of the intake stroke can be increased, so only one valve early closing control is executed as shown in FIG. Compared to the case, both the swirl ratio and the flow coefficient can be increased, that is, the trade-off relationship between the swirl ratio and the flow coefficient can be improved.

  Next, the both-valve quick-open control + single-valve quick-close control (white triangle mark) and the single-valve quick-open / late-close control (black circle mark) are compared. According to the above setting of the second drive cam 84 and the above setting that uses the second drive cam 84 as the first intake valve 64 disposed in the first intake port (main flow port) 60, When the opening and closing control is executed, the flow from the main flow port (first intake port 60) mainly responsible for the generation of swirl can be increased in the initial stage of the intake stroke, so that the reduction of the flow coefficient is suppressed. , Can effectively strengthen the swirl. For this reason, when the one-valve rapid opening / closing control is executed, as shown in FIG. 9, the swirl ratio and the flow coefficient are further increased as compared with the case where the both-valve early opening control + one-valve early closing control is executed. The trade-off relationship can be improved. In addition, when the one-valve quick-open / slow-close control is executed, the intake valve 66 required to obtain the same swirl ratio is faster than the double-valve quick-open control + one-valve quick-close control. The amount of closure can be reduced, and thereby the swirl can be strengthened efficiently while suppressing an increase in pump loss.

FIG. 10 is a diagram for explaining the priority of control of the intake valves 64 and 66 executed when there is a request to increase the swirl ratio in the second embodiment of the present invention.
A control C shown in FIG. 10A is a double-valve rapid opening control, and a control G shown in FIG. 10B is that only the opening timing of the single valve (first intake valve 64) is set to be higher than the intake top dead center. One-valve rapid opening control (in other words, one-valve slow opening control that delays the opening timing of the second intake valve 66 relative to the opening timing of the first intake valve 64), which is controlled to a value on the advance side. The control H shown in FIG. 10C is a one-valve quick opening / closing control (in other words, a one-valve slow opening / closing control).

  As shown in FIG. 10, the swirl ratio can be more effectively increased in alphabetical order given to each control in FIG. Conversely, the flow coefficient increases as the alphabet becomes younger. Therefore, in this embodiment, when there is a request to increase the swirl ratio, the double valve rapid opening control (control C) is performed in order to efficiently increase the swirl ratio while minimizing deterioration of pump loss (fuel consumption). The highest priority is given to execution. Then, when the required swirl ratio is not reached only by executing the double valve quick opening control, the single valve quick opening control (control G) is executed. Further, when the required swirl ratio is not reached even by the execution of the one-valve quick opening control, the one-valve quick opening / closing control (control H) is executed.

FIG. 11 is a diagram showing the operation region using the control C, G, and H methods shown in FIG. 10 in relation to the torque of the internal combustion engine 10 and the engine speed.
As shown in FIG. 11, in the present embodiment, the double-valve rapid opening control (control C) is executed in the middle load region. Then, since a higher swirl ratio is required in the region where the load and the engine speed are higher than the execution region of the control C, the one-valve rapid opening control (control G) is executed. Furthermore, since a higher swirl ratio is required in the region where the load and the engine speed are higher than those in the execution region of the control G, the one-valve quick opening / closing control (control H) is executed.

Next, with reference to FIG. 12, the specific process in Embodiment 2 of this invention is demonstrated. FIG. 12 is a flowchart of a routine executed by the ECU 50 in the second embodiment in order to realize the above function.
In the routine shown in FIG. 12, it is first determined whether or not there is a request to increase the swirl ratio (step 200). More specifically, for example, it is determined whether there is a request to increase the load of the internal combustion engine 10.

  As a result, when it is determined that there is a request to increase the swirl ratio, the control C is first selected, and the both-valve quick-opening control is executed (step 202). Next, it is determined whether or not the required swirl ratio is achieved by the swirl generated only by the execution of the both-valve rapid opening control (step 204).

  As a result, if it is determined that the required swirl ratio is not achieved only by executing the control C, then the control G is selected, and the one-valve rapid opening control is executed (step 206). More specifically, the switching from the control C to the control G is delayed by the VVT mechanism provided in the intake variable valve mechanism 80 so that the opening timing of the second intake valve 66 is close to the intake top dead center. This is realized by using the one-valve variable function so that the closing timing of the second intake valve 66 is approximately the same as the closing timing of the first intake valve 64 driven by the second drive cam 84.

  Next, it is determined whether or not the required swirl ratio is not achieved by the execution of the control G (step 208). As a result, when it is determined that the required swirl ratio is not achieved even by the execution of the control G, the control H is selected, and the one-valve quick opening / closing control is executed (step 210). More specifically, the switching from the control G to the control H is performed by using the one-valve variable function so that the operating angle of the second intake valve 66 is reduced and the second intake valve 66 is quickly closed. It is realized by.

  According to the routine shown in FIG. 12 described above, when there is a request to increase the swirl ratio, the intake is performed according to the priority order of the two-valve rapid opening control, the one-valve rapid opening control, and the one-valve rapid opening / closing control. The swirl control by the control of the valves 64 and 66 is executed. According to such a method, it is possible to efficiently increase the swirl ratio while suppressing deterioration of pump loss (fuel consumption) to a minimum.

  By the way, in the second embodiment described above, the intake variable valve mechanism 80 that can change the working angle (and lift amount) of both intake valves 64 and 66 by using the both-valve variable function is used. However, the present invention is not limited to such a mechanism, and a mechanism that can continuously change only the operating angle (and the lift amount) of the second intake valve 66 may be used. In addition, the intake variable valve mechanism 80 does not necessarily have a phase variable function like the VVT mechanism. If there is a request to increase the swirl ratio, for example, the intake valve 64, 66 may be able to be opened quickly.

In the second embodiment described above, the “intake valve control means” according to the first aspect of the present invention is realized by the ECU 50 executing the processing of steps 200 and 202 described above.
Further, the “flow rate difference generation control means” in the second aspect of the present invention is realized by the ECU 50 executing the processing of steps 204 to 210.
The second drive cam 84 corresponds to the “one-valve quick-open drive cam” in the seventh invention and the “one-valve quick-open drive cam” in the eleventh invention, respectively.

Embodiment 3 FIG.
Next, Embodiment 3 of the present invention will be described with reference to FIGS.
The system of the present embodiment can be realized by causing the ECU 50 to execute a routine shown in FIG. 16 to be described later together with a routine shown in FIG. 6 using the hardware configuration shown in FIG.

[Influence of intake valve opening timing on THC emissions and fuel consumption at low load]
FIG. 13 is a diagram for explaining the influence of the quick opening control of the intake valves 64 and 66 on the THC emission amount and fuel consumption at the time of low load. More specifically, FIG. 13A shows the relationship between the NOx emission amount and the advance amount of the opening timing of the intake valves 64 and 66 (the closing timing is a constant value), and FIG. This represents the relationship between the actual fuel injection amount and the advance amount of the opening timing of the intake valves 64 and 66 (the closing timing is a constant value).

  FIGS. 13A and 13B show the test results during low-load operation with a constant THC discharge. As shown in FIG. 13A, the NOx emission amount increases as the advance amount of the opening timing of the intake valves 64 and 66 increases. The reason is considered as follows. That is, as the advance amount increases, the swirl becomes stronger, and as a result, the amount of THC discharged increases due to excessive diffusion of the spray in the cylinder. In order to reduce the NOx emission amount, it is preferable to increase the internal EGR gas amount. However, if the internal EGR gas amount is increased, the THC emission amount increases. Therefore, in a test performed with a constant THC discharge amount, the increase in the internal EGR gas amount is more limited as the advance amount is increased and the THC discharge amount is increased. As a result, as a result, As the advance angle increases, the NOx emission increases. It should be noted that, in a situation where the EGR rate is constant, the test results having the same tendency as in FIG. 13A can be obtained even when the vertical axis is the THC emission amount.

  As described above, the amount of THC emission increases as the advance amount of the opening timing of the intake valves 64 and 66 increases. Further, as shown in FIG. 13B, when the advance amount is increased, the actual fuel injection amount is decreased (that is, fuel efficiency is improved). This is because the flow coefficient is improved and the pump loss is reduced by the advance angle of the opening timing of the intake valves 64 and 66. From FIG. 13A and FIG. 13B, it can be said that THC and fuel consumption are in a trade-off relationship with respect to the adjustment of the advance amount of the opening timing of the intake valves 64 and 66.

[Relationship between THC emissions and cooling water temperature]
FIG. 14 is a graph showing the relationship between the THC emission amount and the coolant temperature of the internal combustion engine 10.
As shown in FIG. 14, the amount of THC discharged increases as the coolant temperature decreases. This is because the combustion chamber wall surface temperature is lowered when the cooling water temperature is low. As described above, it can be said that THC and fuel consumption are in a trade-off relationship with respect to adjustment of the advance amount of the opening timing of intake valves 64 and 66. Therefore, in consideration of fuel consumption, it is originally intended to increase the above-mentioned advance amount, but as shown in FIG. 14, the THC emission amount increases as the cooling water temperature decreases as shown in FIG. It will increase.

[Characteristic setting of intake valve control in Embodiment 3]
FIG. 15 is a diagram for explaining characteristic settings for controlling the intake valves 64 and 66 according to the third embodiment of the present invention.
In order to solve the above problem, in this embodiment, as shown in FIG. 15, the advance amount of the opening timing of the intake valves 64 and 66 by the both-valve early opening control is reduced as the cooling water temperature becomes lower. did.

FIG. 16 is a flowchart of a routine executed by the ECU 50 in the third embodiment in order to realize the above function. Note that the routine shown in FIG. 16 is executed in parallel with the routine shown in FIG. 6 in the first embodiment.
In the routine shown in FIG. 16, first, the coolant temperature of the internal combustion engine 10 is acquired according to the output of the water temperature sensor 56 (step 300).

  Next, it is determined whether or not the current cooling water temperature is lower than a predetermined temperature, that is, whether or not the internal combustion engine 10 is in a warm-up process (step 302). As a result, when it is determined that the internal combustion engine 10 is in the warm-up process because the cooling water temperature is lower than the predetermined temperature, the opening timing of the intake valves 64 and 66 is determined according to the cooling water temperature. An advance amount is determined (step 304). More specifically, the ECU 50 stores the relationship as shown in FIG. 15 as a map. With reference to such a map, the intake valve 64 based on the both-valve quick-open control as the coolant temperature decreases. , 66 is adjusted so that the advance amount of the opening timing becomes small.

  According to the routine shown in FIG. 16 described above, the opening timing of the intake valves 64 and 66 is advanced as the cooling water temperature rises while appropriately suppressing the THC discharge amount in a situation where the cooling water temperature is low. By going, it will be possible to improve fuel economy. As described above, even during the warm-up process, by appropriately controlling the opening timing of the intake valves 64 and 66, it is possible to reduce both the THC emission amount and the fuel consumption, which are in a trade-off relationship, to the maximum. .

  In the third embodiment described above, the “cold intake valve control means” according to the seventeenth aspect of the present invention is realized by the ECU 50 executing the processing of steps 300 to 304 described above.

Embodiment 4 FIG.
Next, a fourth embodiment of the present invention will be described with reference to FIGS.
The system according to the present embodiment uses the hardware configuration shown in FIG. 1 (however, the configuration of the first exhaust port 120 (which will be described later with reference to FIG. 17) is different from the configuration shown in FIG. 2). This can be realized by executing the routine shown in FIG. 20 described later together with the routine shown in FIG.

[Configuration of Exhaust Port in Embodiment 4]
FIG. 17 is a diagram for explaining a characteristic configuration of the first exhaust port 120 according to Embodiment 4 of the present invention. In FIG. 17, the second intake port 62 and the second exhaust port 70 are not shown.
In the present embodiment, as shown in FIG. 17, the first exhaust port 120 in which the first exhaust valve 72 is disposed uses the first intake port 60 when the exhaust gas flows backward from the first exhaust port 120. Thus, an exhaust swirl flow (an arrow indicated by a broken line in FIG. 17) in a direction opposite to the intake swirl flow (an arrow indicated by a solid line in FIG. 17) obtained can be generated. More specifically, the first exhaust port 120 is configured as a helical helical port having a characteristic that an exhaust swirl flow in the above direction is generated when the exhaust gas flows backward into the cylinder.

[Effects of intake valve opening timing and exhaust valve closing timing on THC emissions at low load]
FIG. 18 is a diagram for explaining the influence of the early opening control of the intake valves 64 and 66 and the exhaust slow closing control of the exhaust valves 72 and 74 on the THC exhaust amount at the time of low load. Note that the white rhombus marks in FIG. 18 indicate the relationship when only the quick opening control of the intake valves 64 and 66 is executed, and is the same as the relationship shown in FIG.

  As described above with reference to FIG. 13, if the intake valves 64 and 66 are opened earlier than the intake top dead center at low load, it is caused by overdiffusion of the spray in the cylinder accompanying the improvement of the swirl. , THC emissions will increase. By the way, at the time of low load, there is a demand for the closing timing of the intake valves 64 and 66 to be a timing near the intake bottom dead center for the purpose of reducing the THC emission amount by improving the actual compression ratio. However, when a mechanism that changes the opening timing in accordance with the closing timing is used, such as the VVT mechanism provided in the intake variable valve mechanism 40, the early closing control of the intake valves 64 and 66 is performed. If this is done, the intake valves 64 and 66 will inevitably be opened quickly, leading to a deterioration in the THC emission accompanying the improvement in swirl. Further, as described above, the THC discharge amount greatly increases as the cooling water temperature decreases.

[Characteristic setting of exhaust valve control in Embodiment 4]
FIG. 19 is a diagram for explaining characteristic settings for controlling the exhaust valves 72 and 74 according to the fourth embodiment of the present invention.
In order to solve the above problems, in this embodiment, as shown in FIG. 19, the retard amount of the closing timing of the exhaust valves 72 and 74 by the exhaust slow closing control is increased as the cooling water temperature becomes lower. .

  The black circles newly added to the graph shown in FIG. 18 indicate the relationship when the exhaust valve slow closing control is executed together with the quick opening control of the intake valves 64 and 66. As shown in FIG. 18, when the exhaust valve late closing control is executed together with the early opening control of the intake valves 64, 66, the NOx emission amount is increased even if the opening timing advance amount of the intake valves 64, 66 is large. It can be seen that the allowance for increasing (that is, THC emissions) is decreasing. This is due to the following reasons. That is, when the exhaust valve closing control for setting the closing timing to the retarded angle side from the intake top dead center is executed simultaneously with the early opening control of the intake valves 64 and 66, the exhaust gas is discharged from the exhaust port at the initial stage of the intake stroke. It will be introduced into the cylinder. As a result, the flow of the intake air introduced into the cylinder and the backflow of the exhaust gas interfere with each other, so that the swirl is suppressed and the THC emission amount is reduced.

  Furthermore, the first exhaust port 120 having the configuration described with reference to FIG. 17 is provided, so that the generation of the intake swirl flow is more effectively suppressed when the exhaust exhaust closing control is performed at a low load. It is possible to sufficiently suppress the discharge of THC.

  FIG. 20 is a flowchart of a routine executed by the ECU 50 in the fourth embodiment to realize the above function. Note that the routine shown in FIG. 20 is executed in parallel with the routine shown in FIG. 6 in the first embodiment. In FIG. 20, the same steps as those shown in FIG. 16 in the third embodiment are denoted by the same reference numerals, and the description thereof is omitted or simplified.

  In the routine shown in FIG. 20, when it is determined that the internal combustion engine 10 is in the warm-up process based on the output of the water temperature sensor 56 and the cooling water temperature is lower than the predetermined temperature, the cooling water temperature is set. Accordingly, the advance amount of the opening timing of the intake valves 64 and 66 is determined (step 304), and the retard amount of the closing timing of the exhaust valves 72 and 74 is determined according to the cooling water temperature (step 304). 400). More specifically, the ECU 50 stores the relationship as shown in FIG. 19 as a map. With reference to such a map, the exhaust valve 72 by the exhaust exhaust closing control, as the coolant temperature decreases, Adjustment is made so that the amount of retardation of the closing timing of 74 increases.

  According to the routine shown in FIG. 20 described above, not only the advance amount of the opening timing of the intake valves 64, 66 but also the retard amount of the closing timing of the exhaust valves 72, 74 are controlled, so that the cooling water temperature is adjusted. Accordingly, it is possible to appropriately suppress the swirl. Thereby, the discharge of THC can be suitably suppressed when cold.

  Further, according to the control of the present embodiment, the swirl ratio can be controlled only by adjusting the valve opening characteristics of the exhaust valves 72 and 74 without providing a separate new swirl suppression means. That is, the control of the present embodiment is performed by adjusting the closing timing of the intake valves 64 and 66 using a VVT mechanism that changes the opening timing and the closing timing at the same time as the intake variable valve mechanism 40 of the present embodiment. This control is suitable for use in a mechanism for performing the above.

  By the way, in the above-described fourth embodiment, in order to effectively suppress the intake swirl flow when the exhaust slow closing control is performed, the exhaust gas flows backward from the first exhaust port 120 to suppress the intake swirl flow. In this case, an exhaust swirl flow in a direction opposite to the intake swirl flow obtained by using the first intake port 60 can be generated. However, in the present invention, the configuration of the exhaust port provided for effectively suppressing the intake swirl flow when the exhaust late closing control is performed is not limited to this. That is, for example, it may be described with reference to FIG. 21 below.

FIG. 21 is a view for explaining a characteristic configuration of the first exhaust port 130 in a modification of the fourth embodiment of the present invention. In FIG. 21, the second intake port 62 and the second exhaust port 70 are not shown.
In the configuration shown in FIG. 21, the first exhaust port 130 is configured as a straight port formed in a straight line. The first exhaust port 130 is oriented so that when the exhaust gas flows backward from the first exhaust port 130 into the cylinder, the exhaust gas flowing backward collides with the intake swirl flow generated by the first intake port 60. Is arranged in. Even with such a configuration, it is possible to more effectively suppress the generation of the intake swirl flow when exhaust exhaust closing control is performed at a low load, and it is possible to sufficiently suppress the discharge of THC. Further, since it is a straight port, it is possible to adjust the intake swirl flow without deteriorating the exhaust efficiency as compared with the case where a helical port is used.

  In Embodiment 4 described above, the ECU 50 executes the processing of Steps 300, 302, and 400, thereby realizing the “cold exhaust valve control means” according to the eighteenth aspect of the invention.

Embodiment 5 FIG.
Next, a fifth embodiment of the present invention will be described with reference to FIG. 22 and FIG.
The system of the present embodiment can be realized by causing the ECU 50 to execute a routine shown in FIG. 23 described later together with a routine shown in FIG. 6 using the hardware configuration shown in FIG.

[Control of intake valve during transition]
In order to improve the acceleration response during a transition that requires rapid acceleration of the vehicle, the fuel injection amount should be increased as much as possible. However, the fuel injection amount that can be increased during acceleration is limited by the upper limit value of the smoke emission amount. On the other hand, increasing the swirl is effective in reducing smoke emissions. However, in the control using the one-valve early closing control or the swirl control valve, if the swirl is increased, the flow coefficient deteriorates and the amount of fresh air sucked into the cylinder is reduced. As a result, the effect of increasing the fuel injection amount by improving the swirl is diminished.

FIG. 22 is a diagram for explaining the control of the opening timing of intake valves 64 and 66 in the fifth embodiment of the present invention.
In order to solve the above-described problem, in the present embodiment, as shown in FIG. 22, the advance angle of the opening timing of the intake valves 64 and 66 at the time of acceleration according to the differential value of the required torque (change rate of the required torque). The amount (based on the intake top dead center) is determined. More specifically, in the period until the design constraint value (determined by the valve stamp) is reached, both the intake valves 64 and 66 both increase as the differential value of the required torque increases, that is, as the required acceleration increases. The advance amount of the opening time is increased. In a region where the difference between the exhaust pressure and the supercharging pressure is large, if the advance amount of the opening timing of the intake valves 64 and 66 is set too large, the smoke discharge amount may be deteriorated due to an increase in the internal EGR gas amount. is there. For this reason, in such a region, the intake valves 64 and 66 are opened quickly only during acceleration (during transition).

FIG. 23 is a flowchart of a routine that the ECU 50 executes in the fifth embodiment in order to realize the above function. Note that the routine shown in FIG. 23 is executed in parallel with the routine shown in FIG. 6 in the first embodiment.
In the routine shown in FIG. 23, it is first determined whether or not the vehicle is accelerating based on the outputs of the crank angle sensor 52 and the accelerator opening sensor 54 (step 500).

  As a result, if it is determined that the vehicle is accelerating, the differential value of the required torque before and after the current acceleration is calculated based on the accelerator opening information such as the amount and rate of change of the accelerator opening during the current acceleration. (Step 502). Next, the advance amount of the opening timing of the intake valves 64 and 66 is determined according to the calculated differential value of the required torque (step 504). More specifically, the ECU 50 stores the relationship as shown in FIG. 22 as a map, and referring to such a map, the greater the differential value of the required torque, that is, the required acceleration. Adjustment is made so that the advance amount of the opening timing of the intake valves 64 and 66 increases as the speed increases.

  According to the routine shown in FIG. 23 described above, the opening timing of both intake valves 64 and 66 is advanced with respect to the intake top dead center in order to increase the swirl ratio during acceleration. According to the swirl improvement method based on the advance timing of the opening timing of the intake valves 64 and 66, as described above, it is possible to increase the swirl ratio without deteriorating the flow coefficient (see FIG. 3B). . That is, it is possible to improve both the swirl ratio and the intake air amount. For this reason, it is possible to increase the fuel injection amount by improving the swirl ratio without the contradiction of reducing the amount of fresh air sucked into the cylinder. Thereby, the response of the internal combustion engine 10 can be improved satisfactorily during acceleration.

  By the way, in the fifth embodiment described above, the opening timing of both intake valves 64 and 66 is advanced to increase the swirl ratio during acceleration. The opening timing of at least one of the valve 64 and the second intake valve 66 may be advanced during acceleration.

  In the fifth embodiment described above, the “acceleration-time intake valve control means” according to the twenty-first aspect of the present invention is implemented when the ECU 50 executes the processing of steps 500 to 504.

It is a figure for demonstrating the system configuration | structure of Embodiment 1 of this invention. FIG. 2 is a diagram for explaining a specific configuration around an intake / exhaust port in the internal combustion engine shown in FIG. 1. It is a figure for demonstrating the advantage of the swirl improvement by both valve early opening control. It is a figure for demonstrating each control method of the intake / exhaust valve used in Embodiment 1 of this invention. FIG. 5 is a diagram showing an operation region using each method of the controls A to F shown in FIG. 4 in relation to the torque of the internal combustion engine and the engine speed. It is a flowchart of the routine performed in Embodiment 1 of the present invention. It is a perspective view for demonstrating the specific structure of the intake variable valve mechanism in Embodiment 2 of this invention. It is the figure which looked at the intake variable valve mechanism shown in FIG. 7 from the axial direction of the control shaft of the arrow A side in FIG. It is a figure for demonstrating the swirl improvement effect by the one-valve quick opening and closing control implement | achieved by the intake variable valve mechanism in Embodiment 2 of this invention. In Embodiment 2 of this invention, it is a figure for demonstrating the priority of control of the intake valve performed when there exists a request | requirement which raises a swirl ratio. It is the figure which represented the driving | operation area | region using each method of control C, G, and H shown in FIG. 10 with the relationship between the torque of an internal combustion engine, and an engine speed. It is a flowchart of the routine performed in Embodiment 2 of this invention. It is a figure for demonstrating the influence of the quick opening control of the intake valve on THC discharge | emission amount and fuel consumption at the time of low load. It is a figure showing the relationship between the discharge amount of THC and the cooling water temperature of an internal combustion engine. It is a figure for demonstrating the characteristic setting of the control of the intake valve in Embodiment 3 of this invention. It is a flowchart of the routine performed in Embodiment 3 of the present invention. It is a figure for demonstrating the characteristic structure of the 1st exhaust port in Embodiment 4 of this invention. It is a figure for demonstrating the influence of the early opening control of the intake valve and the exhaust valve late closing control of the exhaust valve to the THC discharge amount at the time of low load. It is a figure for demonstrating the characteristic setting of the control of the exhaust valve in Embodiment 4 of this invention. It is a flowchart of the routine performed in Embodiment 4 of this invention. It is a figure for demonstrating the characteristic structure of the 1st exhaust port in the modification of Embodiment 4 of this invention. It is a figure for demonstrating control of the opening timing of the intake valve in Embodiment 5 of this invention. It is a flowchart of the routine performed in Embodiment 5 of this invention.

Explanation of symbols

10 Internal combustion engine 18 Exhaust passage 26 Intake passage 40, 80 Intake variable valve mechanism 42 Exhaust variable valve mechanism 50 ECU (Electronic Control Unit)
60 First intake port (mainstream port)
62 Second intake port (dependent port)
64 First intake valve 66 Second intake valves 68, 120, 130 First exhaust port 70 Second exhaust port 72 First exhaust valve 74 Second exhaust valve 82 First drive cam 84 Second drive cam

Claims (21)

An intake port having at least one port having a characteristic that the swirl ratio increases as the amount of air passing through the interior increases;
An intake variable valve mechanism capable of changing a valve opening characteristic of an intake valve disposed in the intake port;
An intake valve control means for performing all-valve quick opening control for advancing the opening timing of all intake valves arranged in the same cylinder when there is a demand to increase the swirl ratio;
A control device for an internal combustion engine, comprising: At least two intake valves are arranged in the cylinder,
The intake port includes a first intake port in which one of the at least two intake valves is disposed, and a second intake port in which the other of the at least two intake valves is disposed,
The intake variable valve mechanism is a mechanism capable of imparting different valve characteristics between the first intake port and the second intake port,
When the intake valve control means determines that the required swirl ratio is not reached even by the execution of the full-valve rapid opening control, the intake valve control unit is configured to perform intake air between the first intake port and the second intake port. 2. A control apparatus for an internal combustion engine according to claim 1, further comprising flow rate difference generation control means for performing flow rate difference generation control for controlling the intake variable valve mechanism so that a flow rate difference is generated in the intake air flowing in the port. .   The flow rate difference generation control is a one-valve quickening method in which the closing timing of either one of the intake valve arranged in the first intake port and the intake valve arranged in the second intake port is advanced compared to the other closing timing. 3. The control apparatus for an internal combustion engine according to claim 2, wherein the control is a closing control.   The flow rate difference generation control is a one-valve delay that delays the opening timing of one of the intake valve arranged in the first intake port and the intake valve arranged in the second intake port compared to the opening timing of the other. 3. The control apparatus for an internal combustion engine according to claim 2, wherein the control is an opening control.   When the flow rate difference generation control means determines that the required swirl ratio is not reached even by the execution of the flow rate difference generation control, the flow rate difference generation control means controls the intake valve and the second intake port arranged in the first intake port. The intake valve is arranged so that the opening timing of one of the arranged intake valves is delayed compared to the opening timing of the other, and the closing timing of the one intake valve is advanced compared to the closing timing of the other intake valve. The control apparatus for an internal combustion engine according to any one of claims 2 to 4, wherein one-valve slow-opening and early-closing control for controlling the variable valve mechanism is performed.   Of the first intake port and the second intake port, the intake port on which the other intake valve is disposed is a port having higher swirl generation capability than the other intake port. The control apparatus for an internal combustion engine according to claim 4 or 5. The intake variable valve mechanism is a drive cam for driving the other intake valve disposed in the port having a high swirl generation capability, and the opening timing of the one intake valve is compared with the opening timing of the other intake valve. A one-valve quick-open drive cam used to drive the other intake valve independently of the one intake valve when the intake variable valve mechanism is controlled to be delayed.
7. The one-valve quick-open drive cam is fixed to a camshaft at a position on an advance side with respect to a drive cam used to drive the one intake valve. Control device for internal combustion engine.   The flow rate difference generation control is a one-valve speed-up method in which the opening timing of either one of the intake valve arranged in the first intake port and the intake valve arranged in the second intake port is advanced compared to the opening timing of the other. 3. The control apparatus for an internal combustion engine according to claim 2, wherein the control is an opening control.   When the flow rate difference generation control means determines that the required swirl ratio is not reached even by the execution of the first flow rate difference generation control, the intake valve disposed in the first intake port and the second intake air The opening timing of one of the intake valves arranged in the port is advanced compared to the opening timing of the other, and the closing timing of the one intake valve is delayed compared to the closing timing of the other intake valve. 9. The control apparatus for an internal combustion engine according to claim 8, wherein a one-valve quick opening and closing control for controlling the intake variable valve mechanism is performed.   Of the first intake port and the second intake port, the intake port on which the one intake valve is disposed is a port having a higher swirl generation capability than the other intake port. The control apparatus for an internal combustion engine according to claim 8 or 9. The intake variable valve mechanism, as a cam for driving the one intake valve arranged in the port having a high swirl generation capability, compares the opening timing of the one intake valve with the opening timing of the other intake valve. A one-valve quick-open drive cam used to drive the one intake valve independently of the other intake valve when controlling the intake variable valve mechanism so as to accelerate
11. The one-valve quick-open drive cam is fixed to a camshaft at a position on the advance side with respect to a drive cam used to drive the other intake valve. Control device for internal combustion engine.   2. The control apparatus for an internal combustion engine according to claim 1, wherein the intake valve control means performs the all-valve quick opening control when an operating range of the internal combustion engine is in a middle and high load range. The internal combustion engine further includes an exhaust variable valve mechanism that can change a valve opening characteristic of the exhaust valve,
The intake valve control means opens the opening timing of all intake valves arranged in the cylinder when the operating range of the internal combustion engine is in the low load range, and opens when the operating range of the internal combustion engine is in the medium load range. Perform full valve slow-open control to control the value to the retard side of the timing,
The control apparatus for an internal combustion engine sets a closing timing of the exhaust valve when the operating region of the internal combustion engine is in a low load region, and is on a more retarded side than a closing timing when the operating region of the internal combustion engine is in a medium load region. The control apparatus for an internal combustion engine according to claim 1 or 12, further comprising exhaust valve control means for performing exhaust slow closing control to be controlled to a value.   The intake valve control means determines that the discharge of unburned HC cannot be sufficiently suppressed even by the execution of the all valve slow open control and the exhaust slow close control when the operating range of the internal combustion engine is in a low load range. 14. The control apparatus for an internal combustion engine according to claim 13, wherein the closing timing of all the intake valves is controlled to a timing in the vicinity of the intake bottom dead center. The intake valve control means performs the all-valve quick-open control when the operating range of the internal combustion engine is in a high rotation and high load range,
The intake valve control means further includes any one of an intake valve disposed in the first intake port and an intake valve disposed in the second intake port when the operating region of the internal combustion engine is in a high rotation / high load region. The control apparatus for an internal combustion engine according to any one of claims 1 and 12 to 14, wherein one-valve early closing control is performed in which one closing timing is advanced compared to the other closing timing. The control apparatus for an internal combustion engine further includes intake pressure acquisition means for acquiring pressure in the intake passage of the internal combustion engine, and exhaust pressure acquisition means for acquiring pressure in the exhaust passage of the internal combustion engine,
16. The internal combustion engine according to claim 15, wherein the intake valve control means decreases the advance amount of the intake valve opening timing by the full valve early opening control as the difference between the exhaust pressure and the intake pressure increases. Engine control device.   The intake valve control means includes a cold-time intake valve control means for reducing the advance amount of the opening timing of the intake valve by the full valve early opening control as the cooling water temperature of the internal combustion engine becomes lower. The control device for an internal combustion engine according to claim 1. The internal combustion engine further includes an exhaust variable valve mechanism that can change a valve opening characteristic of the exhaust valve,
The control apparatus for an internal combustion engine further comprises cold exhaust valve control means for increasing a retard amount of the closing timing of the exhaust valve as the cooling water temperature of the internal combustion engine becomes lower. Control device for internal combustion engine.   The exhaust port in which the exhaust valve is disposed has a direction opposite to an intake swirl flow generated by at least one of the first intake port and the second intake port when exhaust gas flows backward from the exhaust port into the cylinder. 19. The control device for an internal combustion engine according to claim 18, wherein the control device is a port for generating an exhaust swirl flow.   The exhaust port in which the exhaust valve is disposed is a swirl in which at least one of the first intake port and the second intake port generates an exhaust gas that flows backward when the exhaust gas flows backward from the exhaust port into the cylinder. 19. The control device for an internal combustion engine according to claim 18, wherein the control port is a straight port arranged in such a direction as to collide with a flow.   2. The control of an internal combustion engine according to claim 1, wherein the intake valve control means includes an acceleration intake valve control means for advancing an opening timing of the at least one intake valve in order to increase a swirl ratio during acceleration. apparatus.

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