JP2009299623A - Control device for internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve acceleration performance and reduce smoke during acceleration in an internal combustion engine provided with a variable swirl device in relation to a control device for an internal combustion engine. <P>SOLUTION: This control device for the internal combustion engine is provided with the variable swirl device capable of switching between a low swirl mode and a high swirl mode, an intake variable valve gear capable of varying intake valve open timing, a stable operation control means setting the variable swirl device to the high swirl mode in steady state operation in a prescribed operation area, and a transient operation control means setting intake valve open timing earlier than intake valve open timing in the prescribed operation area and making the variable swirl device go through the low swirl mode in transient operation shifting from area where load is lower than that in the prescribed operation area to the prescribed operation area. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

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

  2. Description of the Related Art A variable swirl device that changes the strength of a swirl formed in a cylinder of an internal combustion engine is known. For example, in Japanese Patent Application Laid-Open No. 2004-100555, a case where the lift amount of the first intake valve and the second intake valve is made the same and a case where the lift amount of the first intake valve and the second intake valve are made different are selected. A variable valve mechanism that is practically feasible is disclosed. According to this mechanism, by making the lift amounts of the first intake valve and the second intake valve different, the intake flow rate from each valve can be made different to generate a strong swirl in the cylinder.

JP 2004-1000055 A JP 2004-36566 A

  In an internal combustion engine equipped with a variable swirl device, the mixing of fuel and air can be promoted and smoke can be suppressed by strengthening the swirl during medium and high load operation where smoke is likely to occur.

  However, when the swirl is strengthened by the variable swirl device during the acceleration from the low load region to the medium / high load region, there are the following problems. At the time of acceleration, fresh air tends to be insufficient due to a delay in supercharging response (so-called turbo lag) and a delay in cutting off EGR gas. When the swirl is strengthened in such a state, the flow coefficient is lowered, so that the amount of fresh air is further reduced. As a result, acceleration failure and smoke increase are likely to occur.

  The present invention has been made in view of the above points, and provides an internal combustion engine control device capable of improving acceleration performance and reducing smoke during acceleration in an internal combustion engine including a variable swirl device. With the goal.

In order to achieve the above object, a first invention is a control device for an internal combustion engine,
A variable swirl device that can be switched between a low swirl mode and a high swirl mode that makes a swirl formed in a cylinder of an internal combustion engine stronger than the low swirl mode;
An intake variable valve operating device that makes the intake valve opening timing variable;
When performing steady operation in a predetermined operation region, steady operation control means for setting the variable swirl device to the high swirl mode;
At the time of transient operation that shifts from a lower load region to the predetermined operation region than the predetermined operation region, the intake valve opening timing is made earlier than the intake valve opening timing in the predetermined operation region, and the variable swirl device is moved to the low swirl device. A transient operation control means for passing through the mode, and
It is characterized by providing.

The second invention is the first invention, wherein
The transient operation control means, following the state, makes the intake valve opening timing earlier than the intake valve opening timing in the predetermined operation region, and passes the variable swirl device in the high swirl mode. Features.

The third invention is the first or second invention, wherein
An exhaust variable valve operating device that makes the exhaust valve closing timing variable;
A sudden acceleration request detecting means for detecting a sudden acceleration request;
When the sudden acceleration request is detected, the intake valve opening timing is made earlier than usual and the exhaust valve closing timing is made later than usual so that the intake valve opening period and the exhaust valve opening time are reduced. A valve overlap enlarging means for enlarging the overlap with the valve opening period;
With
The valve overlap enlarging means is characterized by firstly opening the intake valve opening timing and then delaying the exhaust valve closing timing.

  According to the first invention, when switching from the low swirl mode to the high swirl mode at the time of acceleration, it is possible to pass through a state in which the intake valve opening timing is temporarily advanced in the low swirl mode. Thereby, the amount of fresh air and the swirl ratio during acceleration can be increased, and acceleration performance can be improved and smoke can be suppressed.

  According to the second invention, when switching from the low swirl mode to the high swirl mode at the time of acceleration, following the state where the intake valve opening timing is temporarily advanced in the low swirl mode, the intake valve opening timing is advanced The high swirl mode can be passed through. As a result, the swirl ratio can be further increased while securing a sufficient amount of fresh air during acceleration. For this reason, generation | occurrence | production of smoke can be suppressed more reliably.

  According to the third invention, when the valve overlap is provided in order to use the exhaust pulsation during the rapid acceleration, the intake valve opening timing can be first advanced. Thereby, since a fresh air quantity and a swirl ratio can be increased, acceleration performance improvement and smoke suppression can be aimed at. Further, at this stage, the valve overlap period is small, and the lifts of the intake valve and the exhaust valve during the valve overlap period are also small. Therefore, even if the intake pressure is lower than the exhaust pressure due to a delay in response to supercharging, it is possible to reliably suppress the exhaust gas from returning from the exhaust port to the intake port. After that, by delaying the exhaust valve closing timing, a sufficient valve overlap period can be provided, and the volume efficiency can be improved by utilizing the exhaust pulsation. For this reason, excellent rapid acceleration performance can be obtained.

  Embodiments of the present invention will be described below with reference to the drawings. In addition, the same code | symbol is attached | subjected to the element which is common in each figure, and the overlapping description is abbreviate | omitted.

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 a four-cycle diesel engine (internal combustion engine) 10. It is assumed that the diesel engine 10 is mounted on a vehicle and used as a power source. Although the illustrated diesel engine 10 is an in-line four-cylinder type, the number of cylinders and the cylinder arrangement of the diesel engine in the present invention are not particularly limited.

  Each cylinder of the diesel engine 10 is provided with an injector 12 that injects fuel directly into the cylinder. The injectors 12 of each cylinder are connected to a common common rail 14. In the common rail 14, high-pressure fuel pressurized by the supply pump 16 is stored. Then, fuel is supplied from the common rail 14 to each injector 12.

  An exhaust passage 18 of the diesel engine 10 is connected to an exhaust port 22 of each cylinder via an exhaust manifold 20. The diesel engine 10 according to this embodiment includes a turbocharger 24. The exhaust passage 18 is connected to the exhaust turbine of the turbocharger 24.

  An exhaust purification device 26 that purifies exhaust gas is provided in the exhaust passage 18 downstream of the turbocharger 24. As the exhaust purification device 26, for example, one of an oxidation catalyst, a NOx storage reduction type or selective reduction type NOx catalyst, a DPF (Diesel Particulate Filter), a DPNR (Diesel Particulate-NOx-Reduction system), or a combination thereof Etc. can be used.

  An air cleaner 30 is provided near the inlet of the intake passage 28 of the diesel engine 10. The air drawn through the air cleaner 30 is compressed by the intake compressor of the turbocharger 24 and then cooled by the intercooler 32. The intake air that has passed through the intercooler 32 is distributed to the intake port 35 of each cylinder by the intake manifold 34.

  An intake throttle valve 36 is installed between the intercooler 32 and the intake manifold 34 in the intake passage 28. Further, an air flow meter 38 for detecting the amount of intake air is installed in the vicinity of the intake passage 28 downstream of the air cleaner 30.

  One end of an EGR passage 40 is connected to the intake passage 28 in the vicinity of the intake manifold 34. The other end of the EGR passage 40 is connected to the vicinity of the exhaust manifold 20 of the exhaust passage 18. In this system, a part of the exhaust gas (burned gas) can be recirculated to the intake passage 28 through the EGR passage 40, that is, EGR (Exhaust Gas Recirculation) can be performed.

  An EGR cooler 42 for cooling the EGR gas is provided in the middle of the EGR passage 40. An EGR valve 44 is provided downstream of the EGR cooler 42 in the EGR passage 40. By adjusting the opening degree of the EGR valve 44, the EGR amount (EGR rate) can be controlled.

  The system of the present embodiment further includes a supercharging pressure sensor 46 that detects a supercharging pressure, an accelerator position sensor 48 that detects the amount of depression of an accelerator pedal, and an ECU (Electronic Control Unit) 50. The ECU 50 is electrically connected to the various sensors and actuators described above. The ECU 50 controls the operating state of the diesel engine 10 by driving each actuator according to a predetermined program based on the output of each sensor.

  FIG. 2 is a view showing a cross section of one cylinder of the diesel engine 10 in the system shown in FIG. As shown in FIG. 2, a crank angle sensor 62 that detects the rotation angle of the crankshaft 60 is attached in the vicinity of the crankshaft 60 of the diesel engine 10. The ECU 50 can calculate the engine speed based on the signal from the crank angle sensor 62. Further, the ECU 50 calculates an engine load (a fuel injection amount or a required torque) based on a signal from the accelerator position sensor 48, an engine speed, an intake air amount, a supercharging pressure, and the like.

  The diesel engine 10 further includes an intake variable valve operating device 54 that makes the valve opening characteristic of the intake valve 52 variable, and an exhaust variable valve operating device 58 that makes the valve opening characteristic of the exhaust valve 56 variable. The intake variable valve operating device 54 and the exhaust variable valve operating device 58 are each electrically connected to the ECU 50. The exhaust variable valve operating device 58 is an essential configuration in the third embodiment to be described later, but may be omitted in the first and second embodiments. That is, in the first and second embodiments, the valve timing of the exhaust valve 56 may be fixed.

  FIG. 3 is a schematic plan view showing two intake ports 35 provided in one cylinder of the diesel engine 10. As shown in the figure, the diesel engine 10 includes two intake ports 35 per cylinder. In the following description, when it is necessary to distinguish between the two intake ports 35, they are referred to as a first intake port 35a and a second intake port 35b. In the present embodiment, the first intake port 35a is a main flow port, and the second intake port 35b is a sub flow port.

  Each cylinder of the diesel engine 10 is provided with an intake valve 52 that opens and closes the first intake port 35a and an intake valve 52 that opens and closes the second intake port 35b. In the following description, when it is necessary to distinguish between the two intake valves 52, the former is called a first intake valve 52a and the latter is called a second intake valve 52b.

The intake variable valve operating apparatus 54 of this embodiment has the following two functions.
(1) The opening timing of the first intake valve 52a and the second intake valve 52b is changed continuously or stepwise.
(2) Switching between the first operation mode in which the closing timing of the first intake valve 52a and the second intake valve 52b is simultaneously performed and the second operation mode in which the second intake valve 52b is closed earlier than the first intake valve 52a. be able to. That is, in the first operation mode, the operating angle (valve opening period) of the second intake valve 52b is the same as that of the first intake valve 52a. In the second operation mode, the second intake valve 52b and the first intake valve The opening timing of 52a is the same, and the operating angle of the second intake valve 52b is shorter than that of the first intake valve 52a.

  The function (1) described above is, for example, by changing the phase of the camshaft that drives both intake valves 52 to continuously advance or retard their opening and closing timings while maintaining a constant operating angle. It can be realized by a known phase variable mechanism. The function (2) can be realized by a known working angle variable mechanism that varies the working angle of the second intake valve 52b continuously or stepwise.

  The specific configuration of the intake variable valve operating device 54 is not limited to the above, and may be any configuration as long as the functions (1) and (2) can be exhibited. For example, a camshaft that drives the intake valve 52 is rotationally driven by an electric motor so that the intake valve 52 can be opened and closed at any time, or the intake valve 52 can be opened and closed at any time by electromagnetically driving. You may use the apparatus to do.

  In the diesel engine 10 of the present embodiment, the intake variable valve operating device 54 as described above also functions as a variable swirl device that varies the strength of the swirl (swirl flow) formed in the cylinder. In the first operation mode in which the closing timing of the first intake valve 52a and the second intake valve 52b is simultaneous, a strong swirl created by the gas flowing into the cylinder from the first intake port 35a that is the main flow port is generated as a side flow. It is weakened by the flow of gas from the second intake port 35b, which is a port. On the other hand, in the second operation mode in which the second intake valve 52b is closed earlier than the first intake valve 52a, the gas inflow from the second intake port 35b is reduced, so that the gas flow from the first intake port 35a is reduced. The strong swirl you make remains. In this manner, when the intake variable valve operating device 54 is set to the first operation mode, the swirl is low, and when the intake variable valve device 54 is set to the second operation mode, the swirl is high. Therefore, in the following, the first operation mode is referred to as “low swirl mode”, and the second operation mode is referred to as “high swirl mode”.

  When the intake variable valve operating device 54 has a function of changing the operating angle of the second intake valve 52b continuously or in multiple stages, the operating angle of the second intake valve 52b is set in the high swirl mode. The smaller the value, the larger the swirl ratio.

  FIG. 4 is a map showing the relationship between the operating range of the diesel engine 10 and the required swirl ratio. In the present embodiment, as shown in FIG. 4, a higher swirl ratio is required as the engine speed and the engine load increase. In the middle / high load region, by increasing the swirl ratio, the mixing of air and fuel can be promoted, thereby suppressing the generation of smoke.

  In order to realize the required swirl ratio as shown in FIG. 4, the ECU 50 sets the variable intake valve operating device 54 to the low swirl mode when in the steady operation state on the low rotation and low load side from the line C in FIG. When the engine is in a steady operation state on the higher rotation and higher load side than C, control is performed to place the intake variable valve operating device 54 in the high swirl mode.

  When acceleration is performed so as to cross the line C in FIG. 4, for example, when an acceleration request from point A to point B in FIG. 4 occurs, the intake variable valve operating device 54 is moved from the low swirl mode to the high speed. Need to switch to swirl mode. However, when switching from the low swirl mode to the high swirl mode at the time of acceleration from point A to point B, there is a problem that the amount of fresh air supplied into the cylinder of the diesel engine 10 tends to be insufficient.

  When the intake variable valve operating device 54 is switched from the low swirl mode to the high swirl mode, the amount of inflow gas from the second intake port 35b decreases, and the flow coefficient decreases. Further, at the time of acceleration, a delay in response to supercharging (so-called turbo lag) or a delay in cutting off EGR gas occurs. The EGR gas cut-off delay is a phenomenon in which the EGR rate does not suddenly decrease because the EGR gas existing inside the intake manifold 34 continues to flow into the cylinder even after the EGR valve 44 is closed. That is. At the time of acceleration, the amount of fresh air tends to be insufficient due to delay in response of supercharging and delay in turning off EGR gas. For this reason, if the flow coefficient is reduced by switching from the low swirl mode to the high swirl mode, the amount of fresh air is more likely to be insufficient. As a result, there is a negative effect that smoke increases or acceleration performance decreases.

  In order to avoid the shortage of the fresh air amount during acceleration (transient operation) as described above, in the present embodiment, control is performed to advance the opening timing of both intake valves 52 during acceleration. FIG. 5 is a diagram for describing control of intake valve 52 during acceleration in the first embodiment of the present invention. The lower part in FIG. 5 is a diagram showing the change with time of the supercharging pressure during acceleration, and the upper part is a first intake valve 52a (indicated by “IN1”), a second intake valve 52b (indicated by “IN2”), FIG. 6 is a diagram showing a valve opening period of an exhaust valve 56 (indicated by “EX”) by a crank angle.

  Before the start of acceleration (low load region), as shown in FIG. 5 (1), the intake variable valve operating device 54 is in a low swirl mode, and the opening timing of both intake valves 52 is the top dead center (TDC). ) It is said that it is a little before. When an acceleration request is detected based on a signal from the accelerator position sensor 48, the acceleration is started. During acceleration, as shown in (2) of FIG. 5, the opening timing of both intake valves 52 is advanced so that both intake valves 52 open at a timing sufficiently before top dead center while remaining in the low swirl mode. Control is executed.

  When it is recognized that the boost pressure detected by the boost pressure sensor 46 has sufficiently increased, as shown in (3) of FIG. Control for switching to the high swirl mode in which the two intake valves 52b are closed earlier than the first intake valves 52a is executed.

  FIG. 6 is a diagram schematically showing a lift curve of the intake valve 52. The lift curve indicated by the thin line in FIG. 6 is for the normal opening time ((1) in FIG. 5), and the lift curve indicated by the thick line in FIG. 6 is for the case where the opening time is advanced (( 2)). As shown in FIG. 6, when the opening timing of the intake valve 52 is advanced, the valve lift (opening amount) in the initial stage of the intake stroke can be increased. The initial stage of the intake stroke is a period in which the flow rate of the intake gas is large during the intake stroke. Therefore, the magnitude of the valve lift at the initial stage of the intake stroke greatly affects the amount of gas sucked into the cylinder. For this reason, the amount of gas sucked into the cylinder can be sufficiently increased by increasing the lift of the intake valve 52 at the initial stage of the intake stroke. Furthermore, since the lift of the intake valve 52 is increased at the initial stage of the intake stroke, the intake gas flow velocity at the initial stage of the intake stroke is increased, so that the swirl ratio can be increased.

  FIG. 7 is a diagram illustrating the influence of the control for increasing the opening timing of the intake valve 52 on the swirl ratio and the flow coefficient. A black dot in FIG. 7 is a case where the opening timing of the intake valve 52 is a normal timing in the low swirl mode. The solid line extending from the black dot to the upper right indicates the case where the opening timing of the intake valve 52 is advanced while the low swirl mode is maintained. On the other hand, a broken line extending from the black dot to the lower right indicates a case where the mode is switched to the high swirl mode.

  As shown in FIG. 7, when the high swirl mode is set, the swirl ratio can be increased while the flow coefficient decreases. On the other hand, when the opening timing of the intake valve 52 is advanced in the low swirl mode, the flow coefficient can be increased and the swirl ratio can be increased although not as high as the high swirl mode.

  FIG. 8 is a diagram showing changes in swirl ratio, in-cylinder air amount, and EGR rate during acceleration. (1) in FIG. 8 shows the case where acceleration is performed in the high swirl mode. When accelerating in the high swirl mode, the air volume is greatly insufficient due to a decrease in the flow coefficient and a delay in supercharging. For this reason, smoke increases and sufficient acceleration performance cannot be obtained. (2) in FIG. 8 shows a case where acceleration is performed in the low swirl mode while the opening timing of the intake valve 52 is kept normal. In this case, the amount of air is larger than in (1) above, but the swirl ratio is significantly reduced, so that the mixing of air and fuel is insufficient, and smoke increases.

  (3) in FIG. 8 shows the case where the opening timing of the intake valve 52 is advanced to accelerate in the low swirl mode, that is, the control of the present embodiment shown in FIG. 5 is performed. As shown in this figure, according to this embodiment, the amount of air during acceleration (fresh air amount) can be maximized as compared with the cases (1) and (2). In addition, although not as high as the above (1), the swirl ratio can be sufficiently increased. For this reason, generation | occurrence | production of smoke can fully be suppressed. Moreover, since the amount of air can be increased and the amount of fuel injection can be increased accordingly, the acceleration performance can also be improved. Further, since the exhaust energy can be increased by the increase in the air amount and the accompanying increase in the fuel injection amount, the responsiveness of the turbocharger 24 is also improved. That is, since the supercharging delay is reduced, the acceleration performance can be further improved.

  In the present embodiment, the case where the intake variable valve operating apparatus 54 has a function as a variable swirl apparatus has been described. However, in the present invention, the intake variable valve operating apparatus 54 and the variable swirl apparatus are separate. May be. For example, a device that changes the swirl ratio by providing a swirl control valve in the middle of the intake port and opening and closing the swirl control valve may be used as the variable swirl device.

  In the first embodiment described above, the region on the higher rotation and higher load side than the line C in the map shown in FIG. 4 corresponds to the “predetermined operation region” in the first invention. Further, when the ECU 50 switches between the low swirl mode and the high swirl mode of the intake variable valve operating apparatus 54 according to the map shown in FIG. 4 during the steady operation, the “steady operation control means” in the first aspect of the invention is the low swirl mode. By controlling the intake variable valve operating device 54 so as to pass through the state shown in (2) of FIG. 5 during acceleration from the region to the high swirl mode region (during transient operation), the “transient operation” in the first aspect of the invention is described. "Control means" is realized respectively.

Embodiment 2. FIG.
Next, a second embodiment of the present invention will be described with reference to FIG. 9. The description will focus on the differences from the first embodiment described above, and the same matters will be simplified or described. Omitted. The second embodiment is the same as the first embodiment described above except that the control of the intake valve 52 during acceleration is different. FIG. 9 is a diagram for explaining the control of intake valve 52 during acceleration in the second embodiment of the present invention.

  In the present embodiment, as shown in (2) of FIG. 9, when acceleration is started, both the intake valves 52 are opened at a timing sufficiently before top dead center in the low swirl mode. Control for advancing the opening timing of the intake valve 52 is executed. This point is the same as in the first embodiment.

  In this embodiment, after realizing the valve opening characteristic shown in (2) of FIG. 9, the opening timing of both intake valves 52 is kept early as shown in (3) of FIG. Thus, the second intake valve 52b is switched to the high swirl mode that closes earlier than the first intake valve 52a. As a result, the swirl ratio can be further increased while maintaining a high flow coefficient (air amount) due to the early opening timing of both intake valves 52. For this reason, generation | occurrence | production of smoke can be suppressed more reliably.

  Then, when it is recognized that the supercharging pressure detected by the supercharging pressure sensor 46 has sufficiently increased, the opening timing of the both intake valves 52 is restored as shown in (4) of FIG. The state when the acceleration is completed is the same as in the first embodiment.

  In the compression stroke, the gas in the cylinder actually starts to be compressed from the position where the intake valve 52 is closed. For this reason, the substantial compression ratio (hereinafter referred to as “actual compression ratio”) changes according to the closing timing of the intake valve 52. In the state of (3) in FIG. 9, the actual compression ratio is large because the closing timing of the first intake valve 52a is near the bottom dead center (BDC). For this reason, if the state of (3) in FIG. 9 is continuously maintained, the in-cylinder temperature rises and smoke is likely to be generated.

  On the other hand, in this embodiment, as described above, the state is switched to the state of (4) in FIG. 9 with the completion of acceleration. In this state, since the closing timing of the first intake valve 52a is later than the state of (3) in FIG. 9, the actual compression ratio is smaller than the state of (3) in FIG. Therefore, an increase in the in-cylinder temperature can be suppressed, and the generation of smoke can be reliably suppressed.

  In the second embodiment described above, the ECU 50 sequentially passes through the states shown in (2) and (3) of FIG. 9 during acceleration from the low swirl mode region to the high swirl mode region (transient operation). By controlling the intake variable valve operating device 54, the “transient operation control means” in the second aspect of the present invention is realized.

Embodiment 3 FIG.
Next, the third embodiment of the present invention will be described with reference to FIG. 10 to FIG. 13. The description will focus on differences from the above-described embodiment, and the description of the same matters will be simplified. Or omit.

  In the present embodiment, the diesel engine 10 includes an exhaust variable valve operating device 58 that makes the opening timing and closing timing of the exhaust valve 56 variable. In the following description, the variable exhaust valve operating device 58 changes the phase of the camshaft that drives the exhaust valve 56 so that the operating angle (valve opening period) is constant and the opening timing and the closing timing are continuously set. It is assumed that it is composed of a known phase variable mechanism that advances or retards the angle.

  However, the configuration of the variable exhaust valve device 58 is not limited to the above, and a device that rotates the camshaft that drives the exhaust valve 56 by an electric motor so that the exhaust valve 56 can be opened and closed at an arbitrary time. A device that can be opened and closed at any time by electromagnetically driving the exhaust valve 56 may be used.

  In the diesel engine 10 of the present embodiment, when rapid acceleration (full-open acceleration or acceleration close thereto) is performed, control for improving volume efficiency (filling efficiency) by using exhaust pulsation (hereinafter referred to as “volume efficiency improvement control”). Execute). FIG. 10 is a diagram for explaining the volumetric efficiency improvement control. In the volumetric efficiency improvement control, an overlap (hereinafter referred to as “valve overlap”) is provided between the valve opening period of the exhaust valve 56 and the valve opening period of the intake valve 52. That is, the opening timing (IVO) of the intake valve 52 is made earlier than usual, and the closing timing (EVC) of the exhaust valve 56 is made later than usual. The scavenging effect is exhibited by matching the timing of the pulsation valley of the exhaust pressure with the valve overlap period. Thereby, the volumetric efficiency can be improved.

  This will be described in more detail below. The exhaust pressure (exhaust manifold pressure) pulsates (fluctuates) as the exhaust gas is intermittently discharged from the exhaust valve 56 of each cylinder. The waveform of the broken line in FIG. 10 shows the pulsation of the exhaust pressure when the variable exhaust valve device 58 is controlled so that the opening timing (EVO) of the exhaust valve 56 becomes a normal timing. As shown in this waveform, when the opening timing of the exhaust valve 56 is controlled to be a normal timing, the timing of the pulsation valley of the exhaust pressure is the timing before the opening timing of the intake valve 52. It has become.

  On the other hand, the solid line waveform in FIG. 10 shows the pulsation of the exhaust pressure when the variable exhaust valve device 58 is controlled so that the opening timing of the exhaust valve 56 is later than the normal timing. . As the opening timing of the exhaust valve 56 is delayed, the timing at which the exhaust gas is discharged to the exhaust port 22 is delayed, so the waveform (phase) of the exhaust pressure shifts to the right in FIG. Therefore, by appropriately setting the amount by which the opening timing of the exhaust valve 56 is delayed, the timing of the valley of the pulsation of the exhaust pressure can be matched with the valve overlap period as shown in FIG. In the volumetric efficiency improvement control, as described above, the opening timing of the exhaust valve 56 is made slower than usual so that the timing of the valley of the exhaust pressure pulsation coincides with the valve overlap period.

  On the other hand, the alternate long and short dash line in FIG. 10 indicates the intake pressure (intake manifold pressure). As shown in FIG. 10, the intake pressure is substantially constant regardless of the crank angle. For this reason, if the timing of the pulsation valley of the exhaust pressure is made to coincide with the valve overlap period by executing the volumetric efficiency improvement control, the intake pressure can be made higher than the exhaust pressure in the valve overlap period. For this reason, when the intake valve 52 is opened, the burned gas in the cylinder can be quickly expelled to the exhaust port by the fresh air flowing into the cylinder from the intake valve 52. That is, a high scavenging effect can be obtained, and the burned gas in the cylinder can be smoothly and reliably replaced with fresh air. As a result, the residual gas can be sufficiently reduced, and the amount of fresh air filled in the cylinder can be increased accordingly. That is, the volumetric efficiency can be increased and the torque of the diesel engine 10 can be increased.

  However, there is a response delay in turbocharging. For this reason, the intake pressure does not rise sufficiently at the beginning of the rapid acceleration. As a result, the intake pressure may be lower than the exhaust pressure during valve overlap. As a result, the exhaust gas is blown back from the exhaust port 22 to the intake port 35, and the volumetric efficiency is reduced.

  In order to solve this problem, in the present embodiment, when the volumetric efficiency improvement control is started, the opening timing of the intake valve 52 is first advanced without changing the closing timing of the exhaust valve 56. Then, after it is determined that the intake pressure (supercharging pressure) has increased sufficiently, the closing timing of the exhaust valve 56 is delayed.

[Specific Processing in Embodiment 3]
FIG. 11 is a flowchart of a routine executed by the ECU 50 in the present embodiment in order to realize the above function. According to the routine shown in FIG. 11, it is first determined whether or not there is a request for rapid acceleration based on a signal from the accelerator position sensor 48 (step 100).

  If it is determined in step 100 that rapid acceleration is required, the intake variable valve operating apparatus 54 is controlled so that the opening timing of the intake valve 52 is earlier than usual (step 102). FIG. 12 is a diagram showing the valve opening characteristics of the intake valve 52 and the exhaust valve 56 when the processing of step 102 is finished. In the valve opening characteristic shown in FIG. 12, the valve overlap period is small. In the valve overlap period, the intake valve 52 is immediately after opening, and the exhaust valve 56 is immediately before closing. Therefore, the lifts of the intake valve 52 and the exhaust valve 56 are both small. For this reason, even if the intake pressure is lower than the exhaust pressure, the exhaust gas is not easily blown back from the exhaust port 22 to the intake port 35. Therefore, a decrease in volumetric efficiency can be reliably suppressed.

  Furthermore, the valve opening characteristics shown in FIG. 12 have the following advantages. As described in the first embodiment, when the opening timing of the intake valve 52 is advanced, the amount of air can be increased by increasing the flow coefficient, and the swirl ratio can also be increased. For this reason, acceleration performance improvement and smoke suppression can be achieved. Further, since the exhaust energy can be increased by the increase in the air amount and the accompanying increase in the fuel injection amount, the responsiveness of the turbocharger 24 is also improved. That is, since the supercharging delay is reduced, the acceleration performance can be further improved.

  Following the processing in step 102, it is determined whether or not the intake pressure (supercharging pressure) has increased to a value higher than the exhaust pressure in the pulsation valley (step 104). In this step 104, for example, the above determination can be made by comparing the boost pressure detected by the boost pressure sensor 46 with the estimated exhaust pressure calculated based on the operating state of the diesel engine 10. it can.

  If it is determined in step 104 that the intake pressure (supercharging pressure) has become higher than the exhaust pressure in the valley of the pulsation, the variable exhaust valve operating device is set so that the valve timing of the exhaust valve 56 becomes slower than usual. 58 is controlled (step 106). FIG. 13 is a diagram showing the valve opening characteristics of the intake valve 52 and the exhaust valve 56 when the processing of step 106 is finished. In the valve opening characteristics shown in FIG. 13, the valve overlap period is sufficiently secured because the closing timing of the exhaust valve 56 is made later than usual. Further, since the opening timing of the exhaust valve 56 is made later than usual, the valley of the pulsation of the exhaust pressure can be matched with the valve overlap period. Accordingly, the scavenging effect described above is exhibited and the volumetric efficiency can be improved, so that high rapid acceleration performance can be obtained.

  In FIGS. 12 and 13, the second intake valve 52 b is in the high swirl mode in which the operating angle of the second intake valve 52 b is the same as that of the first intake valve 52 a, but the second intake valve 52 b is closed earlier than the first intake valve 52 a. It is good also as a low swirl mode.

  In the above-described third embodiment, the ECU 50 executes the process of step 100, so that the “rapid acceleration request detecting means” in the third aspect of the invention executes the processes of steps 102 to 106. The “valve overlap enlarging means” in the third aspect of the invention is realized.

It is a figure for demonstrating the system configuration | structure of Embodiment 1 of this invention. It is a figure which shows the cross section of one cylinder of the diesel engine in the system shown in FIG. FIG. 2 is a schematic plan view showing two intake ports provided in one cylinder of a diesel engine in the system shown in FIG. 1. It is a map which shows the relationship between the driving | operation area | region of a diesel engine, and a request | requirement swirl ratio. It is a figure for demonstrating control of the intake valve at the time of acceleration in Embodiment 1 of this invention. It is a figure which shows typically the lift curve of an intake valve. It is a figure which shows the influence which the control which makes the opening timing of an intake valve early has on a swirl ratio and a flow coefficient. It is a figure which shows the change of the swirl ratio at the time of acceleration, in-cylinder air amount, and an EGR rate. It is a figure for demonstrating control of the intake valve at the time of acceleration in Embodiment 2 of this invention. It is a figure for demonstrating volumetric efficiency improvement control. It is a flowchart of the routine performed in Embodiment 3 of the present invention. It is a figure for demonstrating control of the intake valve and exhaust valve in Embodiment 3 of this invention. It is a figure for demonstrating control of the intake valve and exhaust valve in Embodiment 3 of this invention.

Explanation of symbols

DESCRIPTION OF SYMBOLS 10 Diesel engine 12 Injector 14 Common rail 18 Exhaust passage 20 Exhaust manifold 22 Exhaust port 24 Turbocharger 26 Exhaust gas purifier 28 Intake passage 34 Intake manifold 35 Intake port 35a First intake port 35b Second intake port 36 Intake throttle valve 38 Air flow Meter 40 EGR passage 44 EGR valve 46 Supercharging pressure sensor 48 Accelerator position sensor 50 ECU
52 Intake valve 52a First intake valve 52b Second intake valve 54 Intake variable valve operating device 56 Exhaust valve 58 Exhaust variable valve operating device 62 Crank angle sensor 64 Piston

Claims (3)

A variable swirl device that can be switched between a low swirl mode and a high swirl mode that makes a swirl formed in a cylinder of an internal combustion engine stronger than the low swirl mode;
An intake variable valve operating device that makes the intake valve opening timing variable;
When performing steady operation in a predetermined operation region, steady operation control means for setting the variable swirl device to the high swirl mode;
At the time of transient operation that shifts from a lower load region to the predetermined operation region than the predetermined operation region, the intake valve opening timing is made earlier than the intake valve opening timing in the predetermined operation region, and the variable swirl device is moved to the low swirl device. A transient operation control means for passing through the mode, and
A control device for an internal combustion engine, comprising:   The transient operation control means, following the state, makes the intake valve opening timing earlier than the intake valve opening timing in the predetermined operation region, and passes the variable swirl device in the high swirl mode. 2. The control apparatus for an internal combustion engine according to claim 1, wherein An exhaust variable valve operating device that makes the exhaust valve closing timing variable;
A sudden acceleration request detecting means for detecting a sudden acceleration request;
When the sudden acceleration request is detected, the intake valve opening timing is made earlier than usual and the exhaust valve closing timing is made later than usual so that the intake valve opening period and the exhaust valve A valve overlap enlarging means for enlarging the overlap with the valve opening period;
With
3. The control apparatus for an internal combustion engine according to claim 1, wherein the valve overlap enlarging means first increases the opening timing of the intake valve and then delays the closing timing of the exhaust valve.

Images