JP2006046128A - Control device for cylinder direct injection type spark ignition internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To achieve both drastic delay of ignition timing and combustion stability, and to realize rise in exhaust gas temperature when an engine is cold and reduction in HC emission amount. <P>SOLUTION: In a warming up completion state where the temperature of cooling water exceeds 80°C, regular stratified charge combustion operation and homogeneous combustion operation are performed. In a cold engine state where the temperature of the cooling water is below 80°C, to promote activation of a catalytic converter, top dead center injection operation mode is performed. In this state, fuel pressure is corrected so as to become higher than that during the stratified charge combustion operation. In addition, injection start timing ITS becomes before a compression top dead center, injection completion timing ITE becomes after the compression top dead center, and fuel injection is performed astride the compression top dead center. Ignition timing ADV becomes timing delayed at 15°CA to 20°CA from the injection start timing ITS after the compression top dead center. Since the compression top dead center becomes stable due to collapse of a large flow and minute turbulence is generated by energy of spray itself, combustion stability is improved. This enables drastic delay of the ignition timing. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to an in-cylinder direct injection spark ignition internal combustion engine that directly injects fuel into a cylinder, and more particularly to control of the injection timing and ignition timing.

Patent Document 1 discloses that when an exhaust purification catalytic converter is in an unwarmed state lower than an activation temperature, fuel is injected during the compression stroke, and the ignition timing is retarded from the compression top dead center. Technology is disclosed.
JP 2001-336467 A

  In order to raise the exhaust gas temperature and reduce HC in order to achieve early activation of the catalyst when the internal combustion engine is cold, it is desirable to retard the ignition timing as much as possible, but if the ignition timing is significantly retarded Since the combustion stability deteriorates, it cannot be retarded from a certain limit determined from the viewpoint of combustion stability. In the above-described conventional technology, it is difficult to ensure stable combustion, particularly under conditions such as cold, the ignition timing delay limit determined from the combustion stability is relatively advanced, and the ignition timing is sufficiently retarded. Cannot be realized.

  The present invention includes a fuel injection valve that directly injects fuel into a cylinder and a spark plug, and fuel pressure variable means that variably controls the fuel pressure supplied to the fuel injection valve, and injects fuel during the compression stroke. In a cylinder direct injection spark ignition internal combustion engine control device that realizes stratified lean combustion at the top dead center when the exhaust gas temperature needs to be raised in a predetermined operating state, for example, when the engine is cold As the injection operation mode, fuel injection is performed in a period that crosses the compression top dead center so that the injection start time is before the compression top dead center and the injection end time is after the compression top dead center, and is delayed from the injection start time. In addition, ignition is performed after compression top dead center, and the fuel pressure is corrected to be higher than that in the stratified lean combustion.

  The fuel pressure in the top dead center injection operation mode desirably increases as the load increases. In particular, it is desirable that the difference between the fuel pressure in the top dead center injection operation mode and the fuel pressure at the same load during stratified lean combustion increases as the load increases.

  FIG. 1 exemplifies a fuel injection period and an ignition timing in the top dead center injection operation mode of the present invention together with a change in in-cylinder pressure. The injection start timing ITS is before the compression top dead center (TDC), and the injection end timing. ITE is after compression top dead center (TDC). The length of the injection period T during that time corresponds to the injection amount. The ignition timing ADV is after compression top dead center (TDC), and is a timing delayed by a predetermined crank angle (for example, 15 ° CA to 20 ° CA) from the injection start timing ITS. This delay period D generally correlates with the distance from the fuel injection valve to the spark plug.

  FIG. 2 shows the piston position change amount and the combustion chamber volume change amount due to the piston stroke in one cycle of the internal combustion engine. As shown in the figure, the amount of change per unit crank angle is the largest near the middle position of the stroke, and is very small near the bottom dead center (BDC) and near the top dead center (TDC). Therefore, in the vicinity of the compression top dead center where the fuel injection is performed in the present invention, the piston position change and volume change are very small, and a stable field that is not affected by the piston movement or the like can be formed.

  In the cylinder, a relatively large gas flow such as a swirl flow or a tumble flow is generated in the intake stroke and remains in the compression stroke. However, a large flow such as a swirl flow or a tumble flow is When the piston reaches near the compression top dead center and the combustion chamber becomes narrow, it collapses rapidly. FIG. 3 shows a change in flow velocity of a large flow in the combustion chamber under various engine speeds. As shown in the figure, a swirl flow or a tumble flow having a strength corresponding to the rotation speed is generated. Collapses rapidly before reaching compression top dead center (360 ° CA). Therefore, in the present invention, the fuel spray injected near the compression top dead center is not moved by a large flow such as a swirl flow or a tumble flow, and always forms a spray in a stable manner on the spark plug. Is possible.

  On the other hand, the energy of a relatively large flow such as the swirl flow or the tumble flow described above transitions to minute turbulence as the flow collapses. Therefore, the minute disturbance in the combustion chamber increases rapidly just before the compression top dead center. FIG. 4 shows the intensity of the minute turbulence caused by the collapse of the flow shown in FIG. 3 as a so-called turbulent flow rate converted to a flow velocity, and as shown in the figure, immediately before the compression top dead center. , Disturbances increase greatly. Such minute disturbances contribute to the activation of the combustion field, and a combustion improving action is obtained.

  In other words, the field in the combustion chamber near the compression top dead center where the fuel is injected does not have a large flow that moves the spray, and there are many minute disturbances that activate the combustion, It is a very stable place against the movement of the piston. In particular, in the present invention, in the top dead center injection operation mode, the fuel pressure is set higher than that in the compression stroke injection for normal stratified lean combustion, and the fuel is injected at a high pressure. In the stable field, an optimum air-fuel mixture can be formed in the vicinity of the spark plug, and the combustion is activated by the disturbance. Therefore, stable combustion is possible with the ignition timing retarded from the compression top dead center, and the retard limit of the ignition timing that is limited in terms of combustion stability is on the retard side. For this reason, the exhaust gas temperature can be significantly increased by a large retardation of the ignition timing, and the HC emission amount is reduced.

  As the load increases, the cylinder pressure at the top dead center increases as the air volume increases.If the fuel pressure is corrected to be higher as the load increases, the spray pressure can be controlled to counter the cylinder pressure. The penetrating force is increased, and an optimum air-fuel mixture can be reliably formed in the vicinity of the spark plug in a high in-cylinder pressure, and turbulence can be generated by spraying. In general, even in the case of compression stroke injection by stratified lean combustion, fuel pressure control is performed so that the fuel pressure increases as the load increases so that the fuel injection period does not become excessively long when the load increases. In order to increase the fuel pressure, it is desirable that the difference between the fuel pressure in the stratified lean combustion and the fuel pressure in the top dead center injection operation mode, that is, the correction amount when compared with the same load, is increased as the load increases.

  According to the present invention, stable combustion can be obtained in a state where the ignition timing is significantly retarded from the compression top dead center. For example, when the internal combustion engine is cold, the exhaust gas temperature is raised and the catalyst is accelerated. Activation can be achieved and HC emissions can be reduced.

  Hereinafter, an embodiment of the present invention will be described in detail with reference to the drawings.

  5 to 7 show an embodiment of a direct injection type spark ignition internal combustion engine to which the present invention is applied. In particular, FIGS. 5 and 6 show the configuration of one cylinder. Indicates the system configuration of the entire organization.

  As shown in FIGS. 5 and 6, a piston 3 is slidably disposed in a cylinder 2 formed in the cylinder block 1, and a cylinder head 4 fixed to the upper surface of the cylinder block 1 and the piston 3 A combustion chamber 5 is formed between them. The cylinder head 4 is formed with an intake port 7 that is opened and closed by an intake valve 6 and an exhaust port 9 that is opened and closed by an exhaust valve 8. A pair of intake valves 6 and a pair of exhaust valves 8 are provided for one cylinder, and an ignition plug 10 is disposed at the center of the ceiling surface of the combustion chamber 5 surrounded by these four valves. . Further, in this embodiment, a partition wall 11 is provided in the intake port 7 so as to partition the intake port 7 into two upper and lower flow paths so that the tumble flow can be strengthened depending on the operation state. A tumble control valve 12 that opens and closes the lower flow path at the upstream end is provided. As can be easily understood by those skilled in the art, the tumble flow is strengthened when the lower flow path is closed by the tumble control valve 12, and the tumble flow is weakened when the tumble control valve 12 is opened. The tumble control valve 12 is not necessarily essential in the present invention, and can be omitted. Alternatively, a known swirl control valve may be provided.

  A fuel injection valve 15 for directly injecting fuel into the cylinder is disposed below the intake port 7 of the cylinder head 4, more specifically at a position between the pair of intake ports 7. The fuel injection valve 15 is arranged so as to inject fuel along a direction orthogonal to a piston pin (not shown) in the plan view, and is arranged so as to be directed obliquely downward in the sectional view of FIG. However, the downward inclination angle is relatively small, that is, the fuel is injected in a direction close to the horizontal.

  On the other hand, the top of the piston 3 has a convex shape along the inclination of the ceiling surface of the combustion chamber 5 that forms a pent roof type, and a concave portion 16 having a substantially rectangular shape in plan view is formed at the center. Yes. The bottom surface of the recess 16 forms an arc surface having a predetermined radius of curvature or a curved surface approximating an arc so as to follow the tumble flow.

  As shown in FIG. 7, the internal combustion engine of this embodiment is, for example, an in-line four-cylinder engine, and a catalytic converter 22 for purifying exhaust gas is provided in an exhaust passage 21 to which an exhaust port 9 of each cylinder is connected. An air-fuel ratio sensor 23 such as an oxygen sensor is disposed on the upstream side. The intake passage 24 to which the intake port 7 of each cylinder is connected is provided with an electronically controlled throttle valve 25 that is opened and closed by a control signal on the inlet side. An exhaust gas recirculation passage 26 is provided between the exhaust passage 21 and the intake air passage 24, and an exhaust gas recirculation control valve 27 is interposed in the middle. Further, the tumble control valves 12 of the respective cylinders are configured to be simultaneously opened and closed by a negative pressure type tumble control actuator 29 that is operated by a suction negative pressure introduced via a solenoid valve 28.

  The fuel injection valve 15 is supplied with fuel that has been adjusted to an appropriate pressure by the fuel pump 31 and the pressure regulator 32 via a fuel gallery 33. Therefore, when the fuel injection valve 15 of each cylinder is opened by the control pulse, an amount of fuel corresponding to the valve opening period is injected. The ignition plug 10 of each cylinder is connected to an ignition coil 34. Here, the pressure regulator 32 is configured to be able to change the fuel pressure of the fuel supplied to the fuel injection valve 15 within a relatively wide range as the fuel pressure varying means.

  The fuel injection timing, injection amount, fuel pressure, ignition timing, etc. of the internal combustion engine are controlled by the control unit 35. The control unit 35 includes a detection signal of an accelerator opening sensor 30 that detects the amount of depression of an accelerator pedal, a detection signal of a crank angle sensor 36, a detection signal of an air-fuel ratio sensor 23, and a detection of a water temperature sensor 37 that detects a cooling water temperature. Signals, etc. are input.

  In the internal combustion engine configured as described above, when the warm-up is completed, for example, when the cooling water temperature exceeds 80 ° C., normal stratified combustion operation and homogeneous combustion operation are performed.

  That is, in a predetermined region on the low speed and low load side, as a normal stratified combustion operation mode, fuel injection is performed at an appropriate time in the compression stroke, with the tumble control valve 12 basically closed. Ignition is performed before the top dead center. In this operation mode, fuel injection always ends before compression top dead center. The fuel injected toward the piston 3 during the compression stroke is collected in the vicinity of the spark plug 10 using a tumble flow swirling along the recess 16 and ignited there. Therefore, stratified combustion with an average air-fuel ratio lean is realized. At this time, the fuel pressure of the fuel injected from the fuel injection valve 15 is increased as the load increases, as shown by the characteristic a in FIG. Control is performed along predetermined characteristics that gradually increase.

  Further, in a predetermined region on the high speed and high load side after the warm-up is completed, fuel injection is performed during the intake stroke under the condition that the tumble control valve 12 is basically opened as a normal homogeneous combustion operation mode. And ignition is performed at the MBT point before the compression top dead center. In this case, the fuel becomes a homogeneous air-fuel mixture in the cylinder and is basically operated near the stoichiometric air-fuel ratio.

  On the other hand, when the cooling water temperature of the internal combustion engine is 80 ° C. or lower, that is, when the warm-up is not completed, the top dead center is used to activate the catalytic converter 22, that is, promote the temperature rise and reduce the HC emission amount. It becomes the injection operation mode. In the top dead center injection operation mode, the fuel pressure controlled by the pressure regulator 32 is sufficiently higher than that in the stratified combustion operation mode described above so that the generation of turbulence due to fuel spray can be obtained. Then, as shown in FIG. 1 described above, the injection start timing ITS is before the compression top dead center (TDC), and the injection end timing ITE is after the compression top dead center (TDC). Is done. The ignition timing ADV is after compression top dead center (TDC), and is ignited at a timing delayed by 15 ° CA to 20 ° CA from the injection start timing ITS. During this delay period, the fuel spray just reaches the vicinity of the spark plug 10 and forms a combustible air-fuel mixture in the vicinity of the spark plug 10, so that ignition combustion is surely performed and stratified combustion is performed. At this time, the fuel injection amount is controlled so that the average air-fuel ratio becomes the stoichiometric air-fuel ratio.

  In this embodiment, the fuel injection timing is controlled so that the injection start timing ITS becomes a predetermined crank angle, and the injection end timing ITE is determined by the injection start timing ITS and the fuel injection amount (injection time). . The injection start timing ITS and the injection end timing ITE are obtained based on the fuel injection amount so that the period before the compression top dead center and the period after the compression top dead center in the fuel injection period are equal. Is also possible.

  As described above, the field in the combustion chamber near the compression top dead center where fuel is injected in this top dead center injection operation mode does not have a large flow that causes the spray to move due to the collapse of the large flow, Along with the collapse of the large flow, there are many minute disturbances that activate the combustion, and the field becomes very stable against the movement of the piston. Then, by performing fuel injection at a high pressure in a stable field where there is no such a large flow, minute turbulence can be positively generated in the cylinder by the energy of the spray itself. Therefore, stable combustion is possible with the ignition timing retarded from the compression top dead center, and the retard limit of the ignition timing that is limited in terms of combustion stability is on the retard side. For this reason, the exhaust gas temperature can be significantly increased by a large retardation of the ignition timing, and the HC emission amount is reduced. FIG. 9 shows the relationship between the minute turbulence generated in the cylinder by the spray itself and the fuel pressure. As shown in the figure, the turbulence is actively generated as the fuel pressure is higher.

  On the other hand, the fuel pressure in the top dead center injection operation mode is controlled with the characteristics as shown in FIG. That is, as a whole, the fuel pressure increases as the load increases, and the correction amount ΔP (in other words, the top dead center injection for the same load when the fuel pressure a in the stratified combustion operation mode is used as a reference). The difference between the fuel pressure b in the operation mode and the fuel pressure a in the stratified combustion operation mode) increases as the load increases. As a reference example, the characteristic c in FIG. 8 shows the fuel pressure characteristic when a certain correction amount is added with reference to the fuel pressure a in the stratified combustion operation mode. The fuel pressure is corrected so as to be higher at the load side.

  FIG. 10 shows a comparison between the change in the in-cylinder pressure when the load is low and the change in the in-cylinder pressure when the load is large. As shown in FIG. The cylinder pressure at the point increases. In contrast to such an increase in the in-cylinder pressure, in this embodiment, the fuel pressure is increased as the load increases as shown by the characteristic b in FIG. 8, so that the penetration force of the spray increases so as to counter the in-cylinder pressure. In addition, an optimum air-fuel mixture can be reliably formed in the vicinity of the spark plug 10 in a high in-cylinder pressure, and disturbance due to the energy of the spray itself is generated more actively.

  In the above embodiment, the fuel injection valve 15 is disposed on the side of the combustion chamber 5 and is configured to inject fuel in a direction close to the horizontal. Of the combustion chamber 5 surrounded by the intake valve 6 and the pair of exhaust valves 8 in the center of the ceiling surface of the combustion chamber 5 so as to inject fuel toward the top surface of the piston 3 at an angle close to vertical. A configuration is also possible.

The characteristic view which showed an example of the fuel-injection period and ignition timing of this invention. The characteristic figure of the piston position change amount and volume change amount during a cycle. The characteristic figure which shows the change in the cycle of a big flow. The characteristic view which shows the change in the cycle of a minute disturbance. Sectional drawing which shows one Example of a direct injection type spark ignition internal combustion engine. FIG. FIG. 2 is a configuration explanatory view showing the system configuration of the entire internal combustion engine. The characteristic view which shows the characteristic of the fuel pressure with respect to the load of this Example. The characteristic view which shows the relationship between the minute disturbance produced | generated by spraying, and fuel pressure. The characteristic view which shows the in-cylinder pressure change by the magnitude of load.

Explanation of symbols

3 ... Piston 5 ... Combustion chamber 10 ... Spark plug 15 ... Fuel injection valve

Claims (4)

  A fuel injection valve for directly injecting fuel into the cylinder and a spark plug, and a fuel pressure variable means for variably controlling the fuel pressure supplied to the fuel injection valve, and stratified lean combustion by injecting fuel during the compression stroke In the control device for a direct injection type spark ignition internal combustion engine that realizes the above, in the predetermined operating state, the fuel injection is performed as the top dead center injection operation mode, and the injection start timing is before the compression top dead center. Is performed in a period straddling the compression top dead center so as to be after the compression top dead center, and ignition is performed after the compression top dead center delayed from the injection start timing, and the fuel pressure is changed from the time of the stratified lean combustion. A control device for an in-cylinder direct injection spark ignition internal combustion engine, wherein   2. The control device for a direct injection type spark ignition internal combustion engine according to claim 1, wherein the fuel pressure in the top dead center injection operation mode increases as the load increases.   The in-cylinder direct injection spark ignition internal combustion engine according to claim 2, wherein the difference between the fuel pressure in the top dead center injection operation mode and the fuel pressure at the same load during stratified lean combustion increases as the load increases. Control device. The in-cylinder direct injection according to any one of claims 1 to 3, wherein the top dead center injection operation mode is executed when a temperature increase of the exhaust gas is requested as a predetermined operation state. -Type spark ignition internal combustion engine control device.

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