JP2003262146A - Control device of spark ignition type direct injection engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To secure ignition stability of mixture gas by maintaining a balance condition of spray penetration of fuel and strength of a tumble flow T without excessively lowering fuel pressure even on a high load side (S1), where an injection volume of the fuel tends to get large and the spray penetration of the fuel gets easily large in a stratified charge combustion area, in a spark ignition type direct jet engine 1, wherein an injector 18 is installed on the peripheral edge of a combustion chamber 6, and the fuel is injected opposing to the tumble flow T of the combustion chamber 6 in a compression stroke of a cylinder 2, so that mixture gas is stratified around an electrode of an ignition plug 16. <P>SOLUTION: The fuel is divided into two and injected by the injector 18 on the high load side (S1) in a stratified charge combustion area during a period after the middle portion of the compression stroke of the cylinder 2. Thereby, the spray penetration of the fuel is lowered, so as to be approximately balanced with the strength of the tumble flow T. On a low load side (S2) of the stratified charge combustion area, the fuel is injected at once by the injector 18 during a period after the middle portion of the compression stroke of the cylinder 2. The exhaust gas is circulated through an EGR passage 33 in the entire stratified charge combustion areas (S1, S2). <P>COPYRIGHT: (C)2003,JPO

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention has a fuel injection valve disposed at the peripheral portion of a combustion chamber in a cylinder, and injects fuel so as to oppose a tumble flow flowing through the combustion chamber during stratified charge combustion operation and ignites the fuel. The present invention relates to a control device for a spark ignition type direct injection engine in which an air-fuel mixture is stratified around electrodes of a plug.

[0002]

2. Description of the Related Art Conventionally, as a spark ignition type direct injection engine of this type, as disclosed in, for example, Japanese Unexamined Patent Publication No. 2000-204954, the tumble force of fuel spray produced by an injector is opposed to each other in the radial direction of a cylinder. There are some that are changed according to the strength of the flow. In this structure, an injector is arranged so that the injection port faces the peripheral portion of the combustion chamber in the cylinder of the engine, a strong tumble flow is generated in the combustion chamber, and the strong tumble flow is collided from approximately the front side. Fuel is injected by the injector. Then, while promoting atomization and vaporization and atomization of the fuel by collision with the tumble flow, the fuel spray is placed on the tumble flow and transported to the electrode side of the spark plug located in the approximate center of the combustion chamber ceiling, The air-fuel mixture is stratified around the electrodes.

Further, in the above-mentioned engine, a throttle valve (tumble control valve: hereinafter referred to as TCV) is provided in the intake passage downstream of the throttle valve, and the TCV is opened and closed by an actuator to enhance the tumble flow in the combustion chamber. The fuel injection pressure from the injector is adjusted to correspond to this, and the penetration force of the fuel spray is balanced with the tumble flow.
By doing so, even if the tumble flow becomes stronger as the engine speed increases, the strength of the fuel spray flow from the injector can be approximately balanced with the strength of the tumble flow in the radial direction of the cylinder regardless of that. As a result, the air-fuel mixture can be transported to the vicinity of the electrodes of the spark plug located substantially in the center of the combustion chamber ceiling.

[0004]

However, in the case of the direct injection engine of the above-mentioned conventional example, since the air-fuel mixture is placed on a strong tumble flow and transported to the spark plug side, the electrodes of the spark plug are ignited by the time of ignition. After that, the air-fuel mixture reaching the vicinity of will pass through the vicinity of the electrode, and since the period during which the air-fuel mixture can be ignited is short, the degree of freedom of ignition timing control becomes extremely low. In addition, there is a problem in ignition stability.

Further, since the fuel spray flow collides with a strong tumble flow from substantially the front, it is inevitable that the fuel spray largely diffuses at the time of this collision and the concentration of the air-fuel mixture becomes low as a whole. Therefore, this also lowers the ignition stability and combustibility of the air-fuel mixture.

On the other hand, the inventors of the present application make the tumble flow flowing through the combustion chamber relatively weak in the compression stroke of the cylinder, and supply the fuel with a relatively small penetrating force that balances the strength of the tumble flow. By injecting, the fuel mixture is gradually decelerated while suppressing the diffusion of the fuel spray so that the air-fuel mixture reaches near the electrode of the spark plug by the ignition timing of the cylinder and is retained there. A method has already been proposed (see Japanese Patent Laid-Open No. 2001-159315).

Also in the direct injection engine of the above-mentioned proposed example, the TCV is arranged in the intake passage, and in the stratified charge combustion region, the TCV is fully closed to generate a tumble flow of a required strength, and this state is maintained. By adjusting the fuel pressure supplied to the injector according to the change in engine rotation speed, the penetration force of the fuel spray is changed to correspond to the strength of the tumble flow, and the equilibrium state of both is maintained. There is.

By the way, in the compression stroke of the cylinder, the pressure of the combustion chamber becomes considerably high, and the fuel spray injected into the combustion chamber is gradually decelerated as a whole while splitting and evaporating due to friction with high density air. As a result, air-fuel mixture becomes mixed and drifts near the electrodes of the spark plug. At this time, while the fuel that first ejects from the injector nozzle quickly loses momentum due to collision with the air wall,
The resistance of the air wall of the subsequent fuel that is continuously ejected after that is eased, and it reaches that far,
This develops fuel spray. Therefore, the penetration force of the fuel spray becomes greater as the injector valve opening time becomes longer, in addition to being influenced by the pressure state of the combustion chamber and the fuel pressure.

From this point of view, on the high load side where the fuel injection amount is relatively large even in the stratified charge combustion region, if the fuel pressure is the same as that on the low load side, the valve opening time of the injector becomes longer, and As the valve opening start timing is advanced, the penetration force of the fuel spray increases, and the equilibrium relationship with the strength of the tumble flow makes the fuel spray slightly superior. In particular, since the tumble flow becomes weak on the low speed side where the intake air velocity is low, the penetration force of the fuel spray tends to be too strong.

On the other hand, it is conceivable to further reduce the fuel pressure on the low speed side and the high load side, but if the required fuel injection amount is to be obtained, the injector valve opening time should be increased in response to the decrease in the fuel pressure. It must be long, and generally lowering the fuel pressure is not preferable from the viewpoint of promoting atomization of the fuel spray.

On the other hand, the direct injection engine of the above proposed example (Japanese Patent Laid-Open No.
No. 001-159315), in which the TCV is fully closed in the entire stratified charge combustion region, the intake resistance due to the TCV becomes considerably large on the high speed side where the intake flow rate is large, and the intake loss cannot be ignored. Therefore, at this time, it is preferable to open the TCV slightly to increase the cross-sectional area of the intake passage. However, if this is done, it is unavoidable that the flow velocity of the intake air decreases and the tumble flow weakens.
As described above, on the high load side of the stratified combustion region where the penetration force of the fuel spray becomes relatively large, there is a risk that the two will not be in balance.

The present invention has been made in view of the above points, and an object of the present invention is to inject fuel so as to oppose a tumble flow in a combustion chamber during stratified charge combustion operation, and to make an electrode of an ignition plug. In a spark-ignition direct injection engine that stratifies the air-fuel mixture around the engine, devises the control procedure of the fuel injection valve when operating on the high load side where the fuel injection amount is relatively large even in the stratified combustion region. In order to maintain a good equilibrium relationship with the tumble flow, the penetration force of the fuel spray is reduced without excessively reducing the fuel pressure.

[0013]

In order to achieve the above-mentioned object, in the solution means of the present invention, the fuel spray flow and the opposing tumble flow are substantially balanced during the stratified charge combustion operation and mixed around the spark plug electrode. In a spark ignition type direct injection engine in which air is stratified, fuel is injected in a plurality of times in a high load side operating state in which a relatively large amount of fuel is injected.

Specifically, according to the first aspect of the invention, while the ignition plug is arranged on the ceiling of the combustion chamber in the cylinder of the engine,
A fuel injection valve is arranged at the peripheral portion of the combustion chamber, and when the engine is in a stratified combustion region, the fuel injection valve injects fuel so as to face the tumble flow in the combustion chamber, and the electrode of the spark plug is provided. It is premised on a control device of a spark ignition type direct injection engine in which the air-fuel mixture is stratified around. Then, when the engine is in the specific region set on the high load side in the stratified charge combustion region, the fuel is injected by the fuel injection valve in a plurality of times in the compression stroke of the cylinder, while the low load of the stratified charge combustion region is injected. On the side, fuel injection control means for collectively injecting fuel in the compression stroke of the cylinder is provided.

With the above construction, first, when the engine is in the stratified charge combustion region, fuel is injected by the fuel injection valve so as to collide from a substantially front side toward the tumble flow flowing in the combustion chamber in the cylinder, and this fuel spray is generated. It is decelerated by the tumble flow and becomes a combustible mixture by the ignition timing of the cylinder and reaches the vicinity of the electrode of the spark plug. that time,
If the engine is in a specific operating region on the relatively high load side, fuel is injected in a plurality of times in the compression stroke of the cylinder by the control of the fuel injection valve by the fuel injection control means (hereinafter, also referred to as divided injection). Say).

When the fuel is dividedly injected in this way, the penetrating force of the fuel spray becomes smaller than that when the fuel is injected all at once. Therefore, for example, the low speed side on the high load side in the stratified charge combustion region is subjected to the specific operation. If the divided injection of fuel is performed at this time as the region (the invention of claim 2), even if the tumble flow becomes weak due to a low intake flow velocity, the fuel pressure does not decrease excessively and the weak pressure is reduced. The penetrating force of the fuel spray can be reduced to match the tumble flow. With this, without impairing the atomization characteristics of the fuel,
By stratifying the air-fuel mixture in the vicinity of the electrodes of the spark plug as desired, good ignition stability can be secured.

Further, for example, tumble strength adjusting means for adjusting the strength of the tumble flow in the combustion chamber by changing the cross-sectional area of the intake passage to the cylinder is provided, and basically the tumble strength is adjusted in the stratified charge combustion region. In order to secure the required tumble strength by reducing the cross-sectional area of the intake passage by means of means, the cross-sectional area of the intake passage in one of the operating regions on the high load side where the output demand to the engine is relatively high. If the tumble strength adjusting means is controlled so that the ratio becomes relatively large and the fuel is dividedly injected (the invention of claim 3), the intake passage cross-sectional area is increased to reduce the intake loss, It becomes easier to secure the output, and even if the strength of the tumble flow decreases due to this, the penetration force of the fuel spray can be reduced to meet this.

More specifically, for example, if the high-speed side of the high-load side of the stratified charge combustion region is set as the specific operation region (the invention of claim 4), the cross-sectional area of the intake passage when the flow rate of intake air is relatively large. Can be increased to reliably prevent an increase in intake loss, and even on the high speed side where it is difficult to secure a time for vaporization and atomization of the fuel spray, it is possible to improve the vaporization and atomization of fuel by the split injection.

At the time of the above-mentioned split injection, it is preferable to inject the fuel in two equal parts within the period after the middle stage of the compression stroke of the cylinder. That is, for example, in a fuel injection valve of a general configuration in which a core valve is operated by an electromagnetic solenoid, there is a minimum limit of the required valve opening time for normal fuel injection operation, and therefore, there is a restriction of split injection. In this case, by performing the injection into two equal parts, in which the valve opening time of one injection operation is the longest, the split injection can be executed up to the low load side where the total injection amount is small. Further, since the time interval at which fuel injection is possible becomes short on the high speed side, split injection in two equal parts is also preferable in this respect.

The specific operation range is not limited to the low load side or the high load side on the high load side in the stratified charge combustion region, but may be set on any of the high load sides. It is also possible to set the entire area on the high load side as the specific operation area.

However, in the split fuel injection, the start timing of the first injection operation is advanced as compared with the case of collective injection, so the fuel spray due to this injection operation spreads relatively large, and Since the concentration of the air-fuel mixture is reduced, the fuel efficiency may be slightly deteriorated. Therefore, it is preferable to perform batch injection on the low load side where the fuel injection amount is relatively small even in the stratified charge combustion region.

According to the invention of claim 5, an exhaust gas recirculation passage for recirculating a part of the exhaust gas to the intake system of the engine, an opening / closing valve for opening / closing the exhaust gas recirculation passage, and the opening / closing valve at least when the engine is in the stratified combustion region. Exhaust gas recirculation control means for opening the valve to recirculate the exhaust gas is provided.

By doing so, at least when the engine is in the stratified charge combustion state, a part of the exhaust gas is recirculated to the intake air and supplied to the combustion chamber, which suppresses the generation of NOx accompanying combustion. By supplying the high-temperature exhaust gas, the vaporization and atomization of the fuel spray is promoted, and the combustibility of the air-fuel mixture is improved.

On the other hand, when the vaporization and atomization of the fuel spray is promoted by the recirculation of the high temperature exhaust gas, the penetration force of the fuel spray is reduced as a result. Therefore, when the ratio of the amount of exhaust gas recirculation to the amount of intake air is equal to or greater than a predetermined value, the split injection of fuel by the fuel injection valve is prohibited even if the engine is in the specific operation region, and the fuel is injected by the fuel injection valve into the compression stroke of the cylinder. It is preferable that the injection is performed all at once (the invention of claim 6).

In this way, the penetration of the fuel spray can be sufficiently reduced by promoting the vaporization and atomization of the fuel spray by the recirculation of a large amount of exhaust gas. It becomes possible to reduce fuel consumption by batch injection of fuel while maintaining the equilibrium relationship with the flow intensity.

According to a seventh aspect of the present invention, on the premise of a control device for a spark ignition type direct injection engine having the same configuration as the first aspect of the present invention, the engine is at least on the low speed side in the high load side in the stratified charge combustion region. At this time, fuel injection control means is provided for injecting fuel by the fuel injection valve so that the injection rate on the injection start side is higher than that on the injection end side in the compression stroke of the cylinder.

With the above construction, when the engine is in the stratified combustion region, the fuel is injected by the fuel injection valve so as to collide with the tumble flow flowing in the combustion chamber in the cylinder substantially from the front, as in the first aspect. Then, the fuel spray is decelerated by the tumble flow and becomes a combustible mixture by the ignition timing of the cylinder and reaches the vicinity of the electrode of the spark plug. At that time, if the engine is at least on the low speed side in the relatively high load side, the fuel injection control unit controls the fuel injection valve so that the injection rate of the fuel in the compression stroke of the cylinder is higher than that in the injection end side. It is injected so that it becomes high.

That is, when the fuel injection valve is opened and fuel is injected in the compression stroke of the cylinder, a relatively large amount of fuel ejected on the injection start side quickly loses momentum due to collision with air. Since the amount of the subsequent fuel (fuel on the injection end side) ejected subsequently to this is relatively small, the development of fuel spray due to the subsequent fuel is suppressed, and even if the injection amount is the same, The penetration force of is relatively small.

As a result, similarly to the second aspect of the invention, the fuel pressure is excessively lowered even in the low-load side on the high load side in the stratified charge combustion region, that is, even when the intake flow velocity is low and the tumble flow is weak. It is possible to reduce the penetration force of the fuel spray to correspond to the weak tumble flow without causing it to stratify the air-fuel mixture in the vicinity of the electrode of the spark plug as desired, and to secure good ignition stability. can do.

In order to perform the injection so that the injection rate on the injection start side becomes higher than that on the injection end side, for example, by changing the substantial opening area during the injection operation by the fuel injection valve, A large opening may be made on the start side of the injection operation, and a small opening may be made on the end side of the injection operation. Alternatively, the fuel injection valve may be intermittently opened and closed a plurality of times, and the valve opening time for each time may be gradually shortened from the injection start side to the injection end side.

Further, as the fuel injection valve, it is preferable to use a fuel injector which is opened and closed by the operation of a piezoelectric element (the invention of claim 8). In this case, compared with a fuel injection valve having a general structure in which a core valve is operated by an electromagnetic solenoid, the degree of freedom in control of injection operation and the accuracy of control are significantly improved, and the injection rate control as described above is performed. Will become more reliable.

[0032]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments of the present invention will be described below with reference to the drawings.

FIG. 1 shows the overall construction of a control device A for a spark ignition type direct injection engine according to the present invention. In the figure, an engine 1 has a cylinder block 3 in which a plurality of cylinders 2, 2, ... (Only one is shown) are provided in series, and a cylinder head 4 arranged on the cylinder block 3. A piston 5 is fitted in each of the cylinders 2 so as to be able to reciprocate in the vertical direction, and a combustion chamber 6 is defined in the cylinder 2 between the crown surface of the piston 5 and the lower surface of the cylinder head 4. It is a thing. Further, a water jacket (not shown) is formed on a side wall portion of the cylinder block 3 surrounding the cylinders 2, 2, ..., Further, a cylinder 2, 3 ,. A crank chamber 7 is formed so as to communicate with the crank chamber 7, and a crank shaft 8 is rotatably disposed inside the crank chamber 7. An electromagnetic crank angle sensor 9 for detecting the rotation angle of the crank shaft 8 is provided on one end side thereof.

The combustion chamber 6 of each of the cylinders 2 is of a so-called pent roof type in which, as shown in an enlarged view in FIG. 2, a roof-like shape in which two slanting surfaces are interlocked with each other is formed in the approximate center of the ceiling. It is a thing. Then, two intake ports 10 and two exhaust ports 11 are opened in each of the two inclined surfaces, and intake and exhaust valves 12, 12 are provided in the respective port openings.
13 are arranged. The intake ports 10 and 10 are
Each of them linearly extends obliquely upward from the combustion chamber 6 and opens independently on one side surface (the right side surface in FIG. 1) of the engine 1 while the two exhaust ports 1
1, 1 and 11 merge into one on the way, extend substantially horizontally, and open to the other side surface (left side surface in FIG. 1) of the engine 1.

The intake valve 12 and the exhaust valve 13 are composed of two cam shafts 14 and 14 which are axially supported inside the cylinder head 4.
Is pressed in the valve axis direction by the valve to be opened, and
By rotating the camshafts 14, 14 in synchronization with the crankshaft 8, the intake valve 12 and the exhaust valve 13 are opened and closed at predetermined timing for each cylinder 2. . Further, the intake side camshaft 14 is provided with a variable valve mechanism 15 having a well-known structure for continuously changing the rotational phase with respect to the crankshaft 8 within a predetermined angle range. Valve 1
The opening / closing operation timing of 2 is changed.

An ignition plug 16 is arranged above the combustion chamber 6 so as to be surrounded by the four intake / exhaust valves 12 and 13. An electrode at the tip of the ignition plug 16 projects from the ceiling of the combustion chamber 6 by a predetermined distance, and an ignition circuit 17 is connected to the base end of the ignition plug 16 to provide a predetermined ignition for each cylinder 2. Spark plug 1 at the timing
6 is energized. On the other hand, a lemon-shaped recess 5a is formed on the crown surface of the piston 5, which is the bottom of the combustion chamber 6.
And the crown surface of the piston 5 outside the concave portion 5a is formed in a mountain shape so as to be substantially parallel to the ceiling portion of the combustion chamber 6 which is opposed thereto.

Further, an injector (fuel injection valve) 18 is arranged on the peripheral portion of the combustion chamber 6 so as to be sandwiched below the two intake ports 10, 10. The injector 18 is a known swirl injector that ejects fuel as a swirl flow from the injection port at the tip and injects it in a hollow cone shape along the direction in which the axis of the injector 18 extends. The fuel spray by the swirl injector 18 has a hollow conical shape under the atmospheric pressure, and the spread angle θ of the fuel spray increases as the atmospheric pressure decreases, while the spread angle θ decreases as the atmospheric pressure increases, and as a whole. Even though it has a conical shape, it becomes a liquid column with a high fuel density in the center. Also, the fuel injection pressure (fuel pressure)
The higher the value, the greater the penetration of fuel spray.

Here, the arrangement relationship between the recess 5a on the crown surface of the piston 5 and the injector 18 will be described in more detail. When viewed along the cylinder center line Z as shown in FIG. The recessed portion 5a is arranged such that its lengthwise direction substantially coincides with the fuel injection direction of the injector 18 (the direction in which the center line F of the fuel spray extends). This means that the direction of intake air flowing from the intake ports 10, 10 into the combustion chamber 6 and the lengthwise direction of the recess 5a substantially coincide with each other, which results in the tumble flow from the intake stroke to the compression stroke of the cylinder 2. T flows toward the injector 18 along the wall surface of the recess 5a, and even if the tumble flow T is relatively weak, it is stably held until the latter stage of the compression stroke of the cylinder 2.

Further, as shown in FIG. 4, when viewed from a direction orthogonal to the cylinder center line Z, the injector 18 has its axial center (corresponding to the fuel spray center line F in this embodiment) of the cylinder 2. It is arranged so as to form a predetermined inclination angle δ (preferably δ = 25 ° to 40 °, about 30 ° in the illustrated example) with respect to the cross section. The spread angle θ of the fuel sprayed by the injector 18 changes depending on the pressure state of the combustion chamber 6 as described above, but in this embodiment, the spray spread angle θ in the compression stroke of the cylinder 2 is within a predetermined range (eg, Uses an injector that fits within θ = approximately 20 ° to approximately 60 °).

With this, that is, the inclination angle δ of the spray center line F and the spray divergence angle θ are set as described above, respectively, the details of which will be described later, but at the time of fuel injection of each cylinder 2 of the engine 1 ( 9), as described above, with respect to the tumble flow T flowing along the concave portion 5a of the crown surface of the piston 5,
Most of the fuel spray can be made to directly face and collide,
The fuel spray can be gradually decelerated by the tumble flow T so that the air-fuel mixture having an appropriate concentration can be retained substantially in the center of the combustion chamber 6.

In order to reliably ignite the air-fuel mixture thus staying, the spark plug 16 has its electrode protruding from the ceiling portion of the combustion chamber 6 along the cylinder center line Z by a predetermined amount. It is arranged. That is, the electrodes of the spark plug 16 are the cylinder block 3 and the cylinder head 4.
The electrode of the spark plug 16 is located closer to the center of the vortex of the tumble flow T from the injection of fuel to the ignition, and the mixture is formed around it. Will be kept in a state of being easily retained.

As described above, all of the injectors 18, 18, ... Arranged for each cylinder 2 are the common fuel distribution pipe 1.
The high-pressure fuel supplied from the fuel supply system 20 is connected to each cylinder 2 through the fuel distribution pipe 19. More specifically, the fuel supply system 20 of this engine is configured, for example, as shown in FIG. 5, and a fuel passage 20b that connects the fuel distribution pipe 19 and the fuel tank 20a.
Low-pressure fuel pump 20 from the upstream side to the downstream side of the
c, a low pressure regulator 20d, a fuel filter 20e, a high pressure fuel pump 20f and a high pressure regulator 20g are arranged in this order. The high-pressure fuel pump 20f and the high-pressure regulator 20g are connected to the fuel tank 20a side by a return passage 20h, and the low-pressure regulator 20i for adjusting the pressure state of the fuel to be returned to the fuel tank 20a side is provided in the return passage 20h. Has been done.
Further, the fuel distribution pipe 19 is provided with a fuel pressure sensor 19a for detecting the pressure of the fuel inside, that is, the pressure of the fuel supplied to the injector 18.

In the fuel supply system 20, the fuel sucked from the fuel tank 20a by the low pressure fuel pump 20c is regulated by the low pressure regulator 20d, filtered by the fuel filter 20e, and sent to the high pressure fuel pump 20f. It The high-pressure fuel pump 20f further boosts the pressure of the fuel, and a part of the fuel is boosted by the high-pressure regulator 2
The flow rate is adjusted by 0 g and supplied to the fuel distribution pipe 19. On the other hand, the surplus fuel is returned to the fuel tank 20a side by the return passage 20h. At this time, the high pressure regulator 2
0 g operates in response to a control signal from the ECU 40, which will be described later, and adjusts the fuel flow rate so that the value detected by the fuel pressure sensor 19a falls within an appropriate range (for example, approximately 3 MPa to approximately 20 MPa). The configuration of the fuel supply system 20 is not limited to the above, and may be a returnless system, for example.

As shown in FIG. 2, an intake passage 21 communicating with the intake ports 10 of each cylinder 2 is connected to one side face of the engine 1. This intake passage 2
Reference numeral 1 is for supplying intake air filtered by an air cleaner (not shown) to the combustion chamber 6 of the engine 1. The hot air for detecting the amount of intake air taken into the engine 1 from the upstream side to the downstream side in order. Wire type air flow sensor 22
An electric throttle valve 23 that throttles the intake passage 21 and a surge tank 24 are provided. The electric throttle valve 23 is not mechanically connected to an accelerator pedal (not shown) but is opened and closed by being driven by an electric drive motor (not shown).

The intake passage 21 on the downstream side of the surge tank 24 is an independent passage branched for each cylinder 2, and the downstream end of each independent passage is further branched into two. It communicates with the intake ports 10, 10. A throttle valve 25 (Tunble Conrol Valve: hereinafter referred to as TCV) that restricts the intake passage 21 in order to adjust the flow velocity of the tumble flow in the combustion chamber 6 is provided on the upstream side of both of the two intake ports 10, 10. The opening / closing operation is performed by, for example, a stepping motor 25a (only shown in FIG. 2).

Each of the TCVs 25 and 25 is formed by cutting out a part of the butterfly valve, and in this embodiment, a part below the valve shaft is cut out. And TC
When V25 is closed, the intake air flows downstream only from the cutout portion, and a strong tumble flow is generated in the combustion chamber 6. On the other hand, as the TCV 25 is opened, the intake air also flows from other than the cutout portion, and the strength of the tumble flow gradually decreases. Therefore, in this embodiment, the TCV 25 and the stepping motor 25a constitute tumble strength adjusting means for changing the cross-sectional area of the intake passage 21 and adjusting the strength of the tumble flow T in the combustion chamber 6.

The shapes of the intake port 10 and the TCV 25 are not limited to those described above, and for example, the intake ports may be so-called common ports that join together on the upstream side. In this case, the TCV may have a butterfly valve having a shape corresponding to the cross-sectional shape of the common port as a base, and may have a shape in which a part of the butterfly valve is cut out as described above. As the tumble strength adjusting means, for example, the variable valve mechanism 15 can change the lift amount of the intake valve 12, and the intake flow velocity is changed by this, thereby adjusting the strength of the tumble flow T. You may do it.

On the other hand, on the other side of the engine 1, the combustion chamber 6
An exhaust passage 26 for discharging burnt gas (exhaust gas) is connected to the exhaust gas. The upstream end portion of the exhaust passage 26 branches for each cylinder 2 and communicates with the exhaust port 11 through the exhaust manifold 2.
7, a linear O2 sensor 28 for detecting the oxygen concentration in the exhaust gas is arranged at the collecting portion of the exhaust manifold 27. The linear O2 sensor 28 is used to detect the air-fuel ratio based on the oxygen concentration in the exhaust gas, and it is possible to obtain a linear output with respect to the oxygen concentration in a predetermined air-fuel ratio range including the theoretical air-fuel ratio. ing.

Further, the exhaust passage 26 on the downstream side of the collecting portion of the exhaust manifold 27 is constituted by an exhaust pipe 29, and the exhaust pipe 29 is provided with HC in the exhaust gas near its stoichiometric air-fuel ratio from its upstream side. , A three-way catalyst 30 for purifying CO, NOx, a lambda O2 sensor 31 for determining the deterioration state of the three-way catalyst 30, and a so-called lean NOx capable of purifying NOx in the exhaust that is leaner than the theoretical air-fuel ratio. The catalyst 32 (including the NOx storage reduction catalyst and the NOx adsorption reduction catalyst) is arranged in order.

Further, an upstream end of an exhaust gas recirculation passage 33 (hereinafter referred to as an EGR passage), which branches off from the exhaust manifold 27 and circulates a part of the exhaust gas to the intake system, is connected to the downstream side of the exhaust manifold 27. There is. A downstream end of the EGR passage 33 is connected to the intake passage 21 between the throttle valve 23 and the surge tank 24, and an electric EGR valve 34 (open / close valve) whose opening degree is adjustable is arranged in the vicinity of the intake passage 21. ing. The opening degree of the EGR valve 34 is adjusted by an electromagnetic solenoid (not shown) or the like, so that the recirculation amount of the exhaust gas flowing through the EGR passage 33 is adjusted.

(Outline of Engine Control) The variable valve mechanism 15, the ignition circuit 17, the injector 18, the fuel supply system 2
0 (high pressure regulator 20g), throttle valve 23, T
The operation of the CV 25, the EGR valve 34, etc. is controlled by an engine control unit 40 (hereinafter referred to as an ECU). On the other hand, the ECU 40 includes at least the crank angle sensor 9, the air flow sensor 22,
Output signals from the linear O2 sensor 28, the lambda O2 sensor 31, etc. are input, and in addition, an accelerator opening sensor 38 for detecting the opening of the accelerator pedal (accelerator operation amount).
And an output signal from a rotation speed sensor 39 that detects an engine rotation speed (rotational speed of the crankshaft 8).

That is, the ECU 40, based on the signals input from the respective sensors, the opening / closing operation timing of the intake / exhaust valves 12 and 13, the ignition timing by the ignition plug 16 of each cylinder 2,
Fuel injection amount by injector 18, injection timing and injection pressure, opening of throttle valve 23, opening of TCV 25, E
The opening degree of the GR valve 34 is controlled respectively. Specifically, for example, as shown in FIG. 6, the stratified combustion region (S
1, S2) are set, and here, as shown in FIGS. 7 (a) and 7 (b), fuel is injected by the injector 18 in the compression stroke of the cylinder 2, and the fuel is injected around the electrode of the spark plug 16. The mixture is burned in a state of being unevenly distributed in layers. Further, at this time, the opening of the throttle valve 23 is made relatively large in order to reduce the intake loss of the engine 1, and the average air-fuel ratio of the combustion chamber 6 at this time is much larger than the theoretical air-fuel ratio. In a lean state (for example, A / F> 3
0).

Further, as a feature of the present invention, which will be described in detail later, on the high load side (S1) in the stratified combustion region (specific operation region), fuel is injected into the cylinder by the injector 18 as shown in FIG. 7 (b). The injection is divided into two equal parts during the middle period of the second compression stroke and thereafter, and on the other hand, on the low load side (S2) in the stratified charge combustion region, the fuel is collectively injected as shown in FIG. 7 (a). I am trying to let you.

On the other hand, except for the stratified combustion region (S1, S2), there is a so-called uniform combustion region (H). Here, the injector 18 mainly injects fuel in the intake stroke of the cylinder 2 to sufficiently mix it with intake air. , A uniform air-fuel mixture is formed over the entire combustion chamber 6 and then burned. This uniform combustion area (H)
In most of the region, the air-fuel ratio of the air-fuel mixture with a uniform fuel injection amount, the opening degree of the throttle valve 23, etc. is approximately the theoretical air-fuel ratio (A / F).
≈14.7), but near full load nearer to the theoretical air-fuel ratio (eg A / F = 12-
14) so that a large output corresponding to a high load can be obtained.

The control of the injector 18 and the throttle valve 23 for switching the combustion state according to the load and the rotation speed of the engine 1 is realized by executing a predetermined control program in the ECU 40, in other words. For example, the ECU 40 includes an injection control unit 40a that controls the fuel injection operation by the injector 18. Then, the injection control unit 40a, when the engine 1 is on the high load side (S1) of the stratified charge combustion region, the injector 1
The fuel is divided into two equal parts for injection in the period after the middle stage of the compression stroke of the cylinder 2 by means of 8, while the fuel is collectively injected on the low load side (S2) of the stratified charge combustion region.

When the engine is warm, the EGR valve 34 is opened in the region shown by hatching in FIG.
A part of the exhaust gas is returned to the intake passage 21 by the passage 33. At this time, the opening degree of the EGR valve 34 is adjusted according to the operating state of the engine 1 such that the exhaust gas recirculation ratio (hereinafter, also referred to as the EGR ratio) decreases as the load increases. This makes it possible to suppress the generation of NOx by the exhaust gas recirculation without impairing the combustion stability of the engine 1. Such control of the EGR valve 34 is also EC
This is realized by executing a predetermined control program in U40, in other words, the ECU 40
Indicates that at least the engine 1 is in the stratified combustion region (S1, S
The exhaust gas recirculation control unit 40b that opens the EGR valve 34 to recirculate the exhaust gas is provided.

As the EGR rate, for example, the ratio of the exhaust gas recirculation amount recirculated to the intake passage 21 by the EGR passage 33 to the fresh air amount may be used. The fresh air is the outside air obtained by removing the recirculation exhaust gas and the fuel gas from the gas sucked into the cylinder 2. Further, when the engine is cold, ensuring combustion stability is given the highest priority, the engine 1 is operated in a uniform combustion state in all operating regions thereof, and the EGR valve 34 is fully closed.

(Formation of Air-fuel Mixture during Stratified Combustion Operation) The feature of the direct injection engine 1 of this embodiment is that the fuel spray injected by the injector 18 in each cylinder 2 in the compression stroke is mainly in the stratified charge combustion state. This is because the tumble flow T of the combustion chamber 6 guides the gas to appropriately stratify the electrode around the spark plug 16. Hereinafter, this point will be described.

First, as shown in FIG. 8, in the intake stroke of the cylinder 2, air is sucked into the combustion chamber 6 from the intake port 10 due to the lowering of the piston 5, and this intake air is introduced from the inner peripheral surface on the exhaust side of the cylinder 2. The tumble flow T is generated by making a large vertical swirl along the crown surface of the piston 5 and further along the inner peripheral surface on the intake side of the cylinder 2. Then, when the cylinder 2 shifts to the compression stroke, the tumble flow T is gradually crushed as the piston 5 rises and becomes compact, but it is appropriate between the ceiling portion of the combustion chamber 6 and the concave portion 5a of the crown surface of the piston 5. Since the space of the shape is left, the tumble flow T is preserved without collapsing until after the middle stage of the compression stroke.

Next, as shown in FIG. 9, at the fuel injection timing of the cylinder 2, fuel is injected by the injector 18 with an appropriate penetration force, and most of this fuel spray is the recess 5a on the crown surface of the piston 5. Tumble flow T flowing along
It collides with the mainstream from the front. Then, the fuel spray advances toward the center of the cylinder 2 while being gradually decelerated by the tumble flow T, and during that time, vaporization and atomization of fuel droplets and mixing with air are promoted, so that the cylinder 2 shown in FIG. At the ignition timing, as shown by hatching in the figure, a mixed air mass forms and stays in the vicinity of the electrode of the spark plug 16.

FIG. 11 shows the calculation result of the deceleration and retention of the fuel spray as described above by applying the computational fluid dynamics (CFD). In the same figure, the flow field of the combustion chamber 6 in the vicinity of the ignition timing of the cylinder 2 is shown. As shown by the thick arrows in the figure, the tumble flow and the fuel spray flow are from the left and right sides of the figure to the center of the cylinder. It can be seen that they are flowing towards each other and they are substantially balanced near the spark plug electrode, which is indicated by the + symbol in the figure. As a result, as shown in FIG. 12, in the vicinity of the ignition timing of the cylinder 2, the air-fuel mixture in an appropriate concentration state is retained around the electrode (ignition position) of the ignition plug 16.

In order to realize the stratification of the air-fuel mixture as described above, the penetrating force of the fuel spray by the injector 18 is approximately balanced with the strength of the opposing tumble flow T, and it is calculated backward from the ignition timing of the cylinder 2. It is necessary to inject fuel at a predetermined timing. Further, it is preferable that both the strength of the tumble flow T and the penetration force of the fuel spray are relatively weak. This is because when the tumble flow T or the fuel spray flow is strong, the fuel spray is largely scattered by the collision even if the two are substantially balanced, and the air-fuel mixture becomes lean as a whole.

Therefore, in the direct injection engine 1 according to this embodiment, the intake ports 10, 10 of each cylinder 2 are straight ports having a large diameter so that the intake flow velocity in the low speed region basically does not become too high. Here, the flow velocity of the intake air is adjusted by restricting the flow of the intake air by the TCV 25 arranged immediately upstream of each intake port 10. Specifically, the engine 1 uses the stratified combustion region (S1,
In step S2), the TCV 25 is fully closed, and the tumble flow T of the required strength is obtained even when the intake flow rate is relatively small.
Is generated, and as shown by the solid line in FIG. 13, the tumble flow T is substantially proportional to the increase in the engine speed ne.
Even if the flow velocity of the cylinder increases, the tumble flow velocity does not become excessively high in the stratified charge combustion region (ne ≦ ne *), and the tumble ratio in the intake stroke of the cylinder 2 is in the range of about 1.1 to about 2.3. Will stay. The graph shown by the solid line in the figure is TCV.
25 is fully closed, and the phantom line shows the TCV 25 when fully open.

As described above, TC depending on the engine speed
The control of the opening degree of V25 is realized by executing a predetermined control program in the ECU 40. In other words, the ECU 40 controls the TCV 25 when the engine 1 is in the stratified charge combustion region (S1, S2). A TCV control unit 40c that increases the flow rate of intake air in a fully closed state is provided.

Further, in the control device A of the direct injection engine 1 of this embodiment, as the strength of the tumble flow T changes according to the engine rotation speed as described above, the strength of the tumble flow T is increased. Fuel injection pressure (fuel pressure) from the injector 18 in order to substantially balance the fuel spray penetration force with the fuel spray penetration force.
Is changed according to the engine speed.
Specifically, the penetration force of the fuel spray may be increased according to the engine speed as shown in FIG. 14, but the penetration force of the fuel spray is not directly proportional to the fuel pressure, and
Since it also changes depending on the valve opening time interval of the injector 18, in actuality, the correspondence between the fuel pressure and the injection pulse width corresponding to a certain fuel injection amount and the penetration force of the fuel spray at that time is experimentally obtained, A table in which the fuel pressure supplied to the injector 18 is recorded in association with the engine rotation speed is created so that the relationship between the rotation speed and the spray penetration force is approximately as shown in FIG. The high pressure regulator 20g of the fuel supply system 20 is controlled according to this table.

Such control of the fuel pressure is realized by executing a predetermined control program in the ECU 40 in accordance with the table electronically stored in the memory. In other words, the ECU 40 controls the stratification of the engine 1 by the ECU 40. When in the combustion region (S1, S2), the fuel supply system 2 is arranged so that the fuel pressure increases in response to the increase in the engine rotation speed.
Fuel pressure control unit 40 for controlling high pressure regulator 20g of 0
d.

(Split injection on the high load side of the stratified charge combustion region)
The feature of the present invention is that, as described above, when the engine 1 is on the high load side (S1) of the stratified charge combustion region, the fuel is divided into two equal parts by the injector 18 in the period after the middle stage of the compression stroke of the cylinder 2. This is due to the injection (see FIG. 7 (b)).

First, the formation process of the fuel spray injected from the nozzle of the swirl injector 18 will be described with reference to FIG.
In the compression stroke of the cylinder 2, the fuel spray S injected from the injection port n of the injector into the high-pressure combustion chamber 6 as a result of friction with high-density air is described below. It splits (atomizes) and evaporates while being mixed with air (vaporization and atomization), then progresses while gradually decelerating and spreads into a hollow conical shape ((a) to (c) in the same figure). Then, after the injection port n is closed, momentum is lost for a while, and the mixture lump C begins to drift in the combustion chamber 6 (Fig. (D),
(e)).

At this time, the fuel initially ejected from the injection port n immediately loses momentum due to collision with the wall of air, while the subsequent fuel is continuously ejected while the fuel is ejected continuously. Fuel spray S without directly colliding with the wall of
Will continue to develop. Particularly, in the case of a swirl injector, a radial vortex v swirling outward is generated at the outer periphery of the tip of the hollow cone-shaped fuel spray S (Fig. (C)),
The vortex v entrains the subsequent fuel and pushes the fuel spray S.

That is, the fuel spray S is developed by the subsequent fuel continuously ejected while the injector nozzle n is opened (Figs. (A) to (c)), and its reaching distance becomes long, The penetration force is affected by the pressure state of the combustion chamber 6 and the fuel injection pressure, and also increases as the injector valve opening time increases.

From this point of view, the stratified combustion region (S1,
Even in S2), the high load side (S
In 1), assuming that the fuel pressure is the same as that on the low load side (S2), the valve opening time of the injector 18 becomes longer, and the valve opening start timing is advanced by that amount, and the combustion chamber 6 is relatively advanced. Since the penetration force of the fuel spray becomes relatively large when the injection starts at a low pressure, the tumble flow T is relatively weak even on the low speed side where the flow velocity of the intake air is low even if the TCV 25 is fully closed. On the other hand, the penetration force of fuel spray tends to be too strong. That is, as in this embodiment, in the direct injection engine 1 in which the fuel-air penetration force and the strength of the tumble flow T are substantially balanced to retain the air-fuel mixture around the electrodes of the spark plug 16, the stratified charge combustion is performed. On the high load side (S1) of the region, on the low speed side, the equilibrium state of the two may be lost.

On the other hand, on the high load side (S
Although it is conceivable to increase the fuel pressure in 1) as compared to the low load side, if the fuel pressure is decreased for the same fuel injection amount, this causes the injector 18 to open the valve for a longer time. Therefore, the penetration force of the fuel spray cannot be lowered as desired, and the fuel pressure is adjusted by the operation of the high-pressure regulator 27 as in this embodiment. It is also difficult to change the fuel pressure without delay. Further, generally, lowering the fuel pressure is not preferable from the viewpoint of promoting atomization of the fuel spray.

Therefore, in this embodiment, the fuel injection is performed on the high load side (S1) of the stratified charge combustion region as described above, so that the penetrating force of the fuel spray is made smaller than in the case of batch injection. ing. That is, FIG.
As schematically shown in FIG. 6, when the fuel is divided into two equal parts and injected in the period after the middle stage of the compression stroke of the cylinder 2, first, the fuel spray Sp by the injection operation of the former stage is collectively injected as shown in FIG. Similarly to the case of (1), it develops while the injection port n is open, and during that time, it gradually diffuses while atomizing and vaporizing and atomizing, and spreads in a hollow conical shape ((a) and (b) in the same figure). The development of the fuel spray Sp in the preceding stage ends once the injection port n of the injector is closed, and thereafter, the fuel spray Sp loses momentum promptly. There, the fuel spray Ss due to the injection operation in the latter stage ((c) in the same figure) catches up ((d) in the same figure) and both are integrated to form the air-fuel mixture C ((e) in the same figure).

As described above, in the case of split injection, the fuel flow due to the injection operation in the latter stage does not contribute to the development of the fuel spray Sp in the former stage, and the fuel spray S due to the injection operation in the latter stage is performed.
Since s directly receives the resistance of the high-density air wall, and further, the effect of propelling the fuel due to the vortex v, which is seen in the case of collective injection, becomes smaller, the two fuel sprays Sp , Ss as a whole, the arrival distance (penetration force) of the fuel spray is considered to be shorter than that in the case of collective injection (Fig. (E)).

Therefore, according to the control device A of the spark ignition type direct injection engine 1 according to this embodiment, when the engine 1 is in the stratified combustion region (S1, S2), the tumble flow T in the combustion chamber 6 of the cylinder 2 is increased. By injecting fuel substantially directly to the fuel cell, and gradually decelerating this fuel spray by the tumble flow T and making it stay near the electrode of the ignition plug 16, the ignition stability is excellent and the degree of freedom of ignition timing control is also high. A high favorable stratified combustion state can be achieved. At that time, in the stratified charge combustion region (S1, S2), the TCV 25 is fully closed so that a certain intake air flow rate can be obtained even when the engine speed is low. As a result, the fuel injection amount is relatively small. On the low load side (S2), the strength of the tumble flow T and the penetrating force of the fuel spray can be balanced as desired.

On the other hand, on the high load side (S1) where the fuel injection amount is relatively large, the penetrating force of the fuel spray is made smaller by the divided injection of fuel than in the case of collective injection,
As a result, the increase in the spray penetration force due to the increase in the fuel injection amount is offset, and the penetration force of the fuel spray is substantially in balance with the strength of the tumble flow T even in the low speed region where the tumble flow T is weak. Can be fastened. That is, even when a large amount of fuel is injected in the low speed region where the tumble flow T becomes weak, the penetrating force of the fuel spray can be reduced to match the weak tumble flow without excessively decreasing the fuel pressure. This makes it possible to stratify the air-fuel mixture in the vicinity of the electrode of the ignition plug 16 as intended without impairing the atomization characteristics of the fuel, thereby ensuring good ignition stability and combustibility.

Moreover, since the fuel is divided into two equal parts in the front stage and the rear stage during the split injection, the valve opening time for each injection becomes the longest with respect to the total injection amount.
Therefore, the swirl injector 18 in which the core valve is operated by the electromagnetic solenoid as in this embodiment
Even if there is a restriction on the minimum valve opening time required for normal fuel injection operation, the split injection can be executed up to the low load side region where the total injection amount is relatively small. Becomes

Further, in this embodiment, as described above, a part of the exhaust gas is recirculated to the intake passage 21 by the EGR passage 33 in the stratified combustion region (S1, S2), whereby NOx is generated by the combustion. In addition to suppressing the above, it is possible to promote the vaporization and atomization of the fuel spray by the high temperature exhaust gas and improve the combustibility.

In the above embodiment, the method of split injection of fuel by the injector 18 may be such that the injection amount of fuel by each injection operation is different, and the number of times of fuel injection is three or more. You can also

(Other Embodiments) The structure of the present invention is not limited to the above-mentioned embodiment, and includes various other structures. That is, in the above-described embodiment, the TCV 25 is always fully closed within the stratified combustion region (S1, S2), but the present invention is not limited to this, and as shown by the phantom line in FIG. It is also possible to configure the TCV control unit 40c so that the TCV 25 is half-opened on the high speed side in S1, S2) and fully closed on the other stratified combustion regions (S1, S2) on the low speed side. In this way, the intake resistance due to the TCV 25 can be reduced on the high speed side where the intake flow rate is high, and the intake loss can be substantially eliminated.

The TCV 25 allows intake air to flow through the notch even in the fully closed state.
The V25 being half open means that the flow rate of intake air is approximately half of the fully closed state and the fully open state.

As described above, when the TCV 25 is in the half-opened state, the flow velocity of the intake air is lower than that in the fully closed state. This can be dealt with by reducing the penetration force by split injection.

Further, as in the above embodiment, the split injection is performed on the high load side (S1) of the stratified charge combustion region, and on the other hand, instead of performing the collective injection on the low load side, the high load side of the stratified charge combustion region (S1) It is also possible to open the TCV 25 more than in the fully closed state in the specific operation region set to), and to perform split injection at this time. In this case, the ECU 40 T
The CV control unit 40c controls the opening degree of the TCV 25 so that the cross-sectional area of the intake passage 21 becomes larger in the specific operation region than in the other stratified combustion regions (S1, S2). Constitutes a means.

Further, in the above-described embodiment, the TCV control unit 40c is set to TCV on the high load side (S1) of the stratified charge combustion region.
25 may be half-opened, while fully closed on the low load side (S2).

Further, in this embodiment, exhaust gas recirculation (EGR) is always performed in the warm stratified combustion region (S1, S2), while the stratified combustion region is irrespective of the EGR rate. However, the present invention is not limited to this, and when the EGR rate (ratio of the exhaust gas recirculation amount to the intake air amount) is equal to or higher than a predetermined value, the engine is in the specific operation region. Even if the divided injection of fuel by the fuel injection valve is prohibited, the injector 18
Therefore, the fuel may be collectively injected in the compression stroke of the cylinder.

In this way, the penetration of the fuel spray can be sufficiently reduced by promoting the vaporization and atomization of the fuel spray by the recirculation of a large amount of exhaust gas. While maintaining the equilibrium relationship with the strength of the flow T, fuel consumption can be reduced by batch injection of fuel.

Furthermore, in the above embodiment, the swirl injector 18 is used as the fuel injection valve for directly injecting the fuel into the combustion chamber 6 in the cylinder 2 of the engine 1, but the invention is not limited to this. For example, a slit-type injector in which a slit is provided at the tip end to form a flat fuel spray, or fuel is injected from a plurality of minute injection ports, and these are integrated to form one fuel spray. The multi-jet type injector described above may be used.

Further, the fuel injection valve is not limited to the one in which the core valve is actuated by the electromagnetic solenoid or the like as in the above embodiment, but an injector of the type opened and closed by a piezoelectric element (piezo element) may be used. it can. In this case, since the injection rate can be controlled, by changing the fuel injection rate during one fuel injection operation instead of the split injection in the above embodiment,
The penetration force of the fuel spray may be adjusted.

For example, as shown in FIG. 18, when the engine 1 is at least on the low speed side in the high load side (S1) of the stratified charge combustion region, fuel is injected by the injector after the middle stage of the compression stroke of the cylinder 2. The injection is performed so that the injection rate on the start side is higher than that on the injection end side. Specifically, as shown in (a) of the same figure, the substantial opening area is changed during one injection operation so that the opening side is wide at the start side of the injection operation and small at the end side of the injection operation. To Alternatively, as shown in FIG. 6B, the injector is intermittently repeated a plurality of times at extremely short time intervals (5 times in the example in the figure,
However, the valve opening time may be gradually shortened as the valve is opened and closed.

When the injection rate is changed in this way, the relatively large amount of fuel (fuel on the injection start side) that is first ejected from the injector when it opens is promptly collided with the air wall. While the momentum is lost, the amount of the subsequent fuel (fuel on the injection end side) ejected subsequently decreases, so that the development of fuel spray due to this subsequent fuel is suppressed and the penetration force of the fuel spray is relatively increased. Will be small. As a result, similar to the split injection in the above-described embodiment, the injection amount of the fuel increases even though the tumble flow T becomes weak due to the low speed side on the high load side in the stratified charge combustion region, that is, the intake flow velocity is particularly low. In some situations, the penetrating force of the fuel spray can be reduced to match the weak tumble flow T without excessively reducing the fuel pressure.

[0091]

As described above, according to the control device for a spark ignition type direct injection engine according to the invention of claim 1, the fuel injection valve is arranged at the peripheral portion of the combustion chamber in the cylinder, and the low speed low load is applied. In the stratified charge combustion region, the fuel is injected by the fuel injection valve so as to face the tumble flow in the combustion chamber to stratify the air-fuel mixture around the electrode of the spark plug. In a specific operating region set on the high load side, the fuel is injected in a plurality of times in the compression stroke of the cylinder, so that even if the fuel injection amount is relatively large, the penetration force of the fuel spray is in the case of collective injection. It can be smaller than.

Therefore, for example, as in the invention of claim 2,
If the low-speed side of the high-load side in the stratified combustion region is set as the specific operation region, the penetration force of the fuel spray and the tumble flow can be maintained without excessively lowering the fuel pressure even in the low-speed region where the intake flow velocity is low and the tumble flow is weak. It is possible to easily maintain the equilibrium state with the strength of the air-fuel mixture and retain the air-fuel mixture in the vicinity of the spark plug electrode as intended, thereby ensuring good ignition stability.

Further, as in the third aspect of the invention, basically, in the stratified charge combustion region, the tumble strength adjusting means is controlled so that the cross-sectional area of the intake passage to the cylinder becomes small. Is controlled so that the cross-sectional area of the intake passage becomes relatively large in any one of the operating regions on the high load side where the output demand is relatively high, and the split injection of fuel is performed. By doing so, the intake loss can be reduced and the output can be easily ensured while maintaining the balance between the penetration force of the fuel spray and the strength during tumble.

More specifically, when the high speed side of the high load side of the stratified charge combustion region is set as the specific operation region as in the invention of claim 4, for example, when the flow rate of intake air is relatively high, the cross sectional area of the intake passage is increased. Can be increased to reliably prevent an increase in intake loss. Moreover, even on the high speed side where it is difficult to secure the time for the vaporization and atomization of the fuel spray, the vaporization and atomization of the fuel can be improved by the divided injection.

According to the fifth aspect of the present invention, by recirculating a part of the exhaust gas to the intake system of the engine in the stratified combustion region, the generation of NOx due to combustion is suppressed, and the vaporization and atomization of the fuel spray is promoted. Thus, the combustibility of the air-fuel mixture can be improved.

In this case, like the invention of claim 6, if the ratio of the exhaust gas recirculation amount is large even in the specific operation region, the split injection is prohibited and the fuel is injected all at once. Since a large amount of exhaust gas recirculation can sufficiently reduce the penetration force of the fuel spray, it is possible to reduce fuel consumption by batch injection while maintaining the equilibrium relationship between the penetration force of the fuel spray and the strength of the tumble flow. .

According to the control apparatus for a spark ignition type direct injection engine according to the invention of claim 7, in the same precondition as the invention of claim 1, the fuel injection valve is used on the low speed side in the high load side in the stratified charge combustion region. By injecting fuel so that the injection rate on the injection start side is greater than that on the injection end side,
Even if the fuel injection amount is large, the penetrating force of the fuel spray can be made relatively small. As a result, similarly to the invention of claim 2, even in the low speed region where the intake air velocity is low and the tumble flow is weak, the air-fuel mixture is retained near the ignition plug electrode without excessively lowering the fuel pressure, which is favorable. Ignition stability can be secured.

According to the eighth aspect of the present invention, by using the fuel injection valve that opens and closes by the operation of the piezoelectric element, the degree of freedom in controlling the injection operation and the accuracy of the control are remarkably increased, and the injection rate can be controlled more reliably. This will allow you to
The effect of the invention of claim 7 can be sufficiently obtained.

[Brief description of drawings]

FIG. 1 is a diagram showing an overall configuration of a control device for a spark ignition direct injection engine according to an embodiment of the present invention.

FIG. 2 is a perspective view showing an arrangement configuration of a piston crown surface, an intake port, an ignition plug, and an injector.

FIG. 3 is an explanatory diagram showing a positional relationship among a concave portion of a piston crown surface, a tumble flow, and a fuel spray when viewed along a cylinder center line.

4 is a view corresponding to FIG. 3 viewed from a direction orthogonal to a cylinder center line.

FIG. 5 is an explanatory diagram schematically showing a configuration of a fuel supply system.

FIG. 6 is a diagram showing an example of a control map in which operating regions for setting the engine to a stratified combustion state or a uniform combustion state are set respectively.

FIG. 7 is an explanatory view schematically showing the fuel injection operation of the injector.

FIG. 8 is a diagram showing a state of a tumble flow generated in a combustion chamber in an intake stroke of a cylinder.

FIG. 9 is a diagram showing a state of fuel spray injected so as to collide with a tumble flow at a fuel injection timing of a cylinder.

FIG. 10 is an explanatory diagram showing a state of an air-fuel mixture that stays near an electrode of a spark plug at an ignition timing of a cylinder.

FIG. 11 is an explanatory diagram showing a result of a CFD analysis in which a tumble flow and a fuel spray are balanced in a combustion chamber.

FIG. 12 is an explanatory diagram showing the concentration distribution of the air-fuel mixture that stays in the vicinity of the electrodes of the spark plug.

FIG. 13 is a graph showing a change in tumble flow rate with respect to a change in engine rotation speed.

FIG. 14 is a graph showing an example of control characteristics for changing the spray penetration force in association with a change in engine rotation speed.

FIG. 15 is an explanatory diagram schematically showing the state of fuel spray when the fuel is divided into two equal parts and injected over time.

FIG. 16 is an explanatory diagram schematically showing a state of fuel spraying when fuel is collectively injected, over time.

FIG. 17 is a view corresponding to FIG. 6 according to another embodiment in which the TCV is opened on the high speed side of the stratified charge combustion region.

FIG. 18 is a view corresponding to FIG. 7 according to another embodiment in which the fuel injection rate is changed.

[Explanation of symbols]

A Spark ignition type direct injection engine control device 1 engine Two cylinder 6 Combustion chamber 16 spark plugs 18 Injector (fuel injection valve) 21 Intake passage 25 TCV (Tumble strength control means) 25a Stepping motor (tumble strength adjusting means) 33 EGR passage (exhaust gas recirculation passage) 34 EGR valve (open / close valve) 40 ECU (engine control unit) 40a Injection control unit (fuel injection control means) 40b EGR control section (exhaust gas recirculation control means) 40c TCV control unit (tumble strength control means) High load side of S1 stratified combustion area (specific operation area) T tumble style

Front page continuation (51) Int.Cl. ⁷ Identification code FI theme code (reference) F02D 41/34 F02D 41/34 H 43/00 301 43/00 301J 301T 301U F02M 25/07 570 F02M 25/07 570A ( 72) Inventor Keiji Araki 3-1-1 Shinchi, Fuchu-cho, Aki-gun, Hiroshima Prefecture Mazda Co., Ltd. (72) Fumihiko Saito 3-3-1 Shinchi, Fuchu-cho, Aki-gun, Hiroshima Prefecture (72) Inventor Maruhara Masashi, 3-1, Shinchi, Fuchu-cho, Aki-gun, Hiroshima Prefecture Mazda Co., Ltd. (72) Inventor Akira Inayama 3-1, Shinchi, Fuchu-cho, Aki-gun, Hiroshima Prefecture (72) Inventor, Yuri Seto Aki-gun, Hiroshima Prefecture Fuchu-cho Shinchi No. 3 Mazda Co., Ltd. F term (reference) 3G023 AA00 AA01 AA18 AB03 AC05 AD02 AD06 AG01 AG03 3G062 BA04 BA05 BA06 BA08 BA09 CA06 DA01 DA02 EA10 ED01 ED04 ED10 FA02 FA05 FA23 GA01 GA04 GA06 GA17 3G084 A08 BA15 BA20 BA21 BA24 CA03 CA04 CA09 DA02 DA04 DA10 DA 11 EC07 FA08 FA10 FA29 FA30 FA33 FA38 3G092 AA01 AA06 AA09 AA10 AA17 BA08 BB06 BB08 BB12 BB13 BB19 DC03 DC06 DC09 DE03S DE09S DG09 EA01 EA11 FA00 FA05 HA15 HA08 HA01 HA23 HA01Z23 KA09 LA03 LA05 LB04 LB06 LC01 MA19 MA26 MA27 PA04Z PB08A PB08Z PD04Z PE01Z PE03Z PF03Z

Claims (8)

[Claims] 1. A fuel injection valve is provided at a peripheral portion of the combustion chamber while an ignition plug is provided at a ceiling portion of a combustion chamber in a cylinder of the engine, and the fuel injection is performed when the engine is in a stratified combustion region. In a control device for a spark ignition type direct injection engine in which fuel is injected by a valve so as to oppose a tumble flow in a combustion chamber to stratify the air-fuel mixture around an electrode of the spark plug, When in a specific operating region set on the high load side in the combustion region, the fuel is injected by the fuel injection valve in a plurality of times in the compression stroke of the cylinder, while the fuel is injected on the low load side in the stratified combustion region. A control device for a spark ignition type direct injection engine, comprising fuel injection control means for performing a single injection in a compression stroke of a cylinder. 2. The specific operation region is set to a low speed side of the high load side in the stratified charge combustion region according to claim 1, and the fuel injection control means controls the fuel injection valve when the engine is in the specific operation region. A control device for a spark ignition type direct injection engine, characterized in that fuel is divided into two equal parts and injected during a period after the middle stage of a compression stroke of a cylinder. 3. The tumble strength adjusting means for changing the cross-sectional area of the intake passage to the cylinder to adjust the strength of the tumble flow in the combustion chamber according to claim 1, and the tumble strength adjusting means when the engine is in a specific operation region. And a tumble strength control means for controlling the tumble strength adjusting means so that the cross-sectional area of the intake passage becomes larger than that in the stratified charge combustion region other than the spark ignition type direct injection engine. Control device. 4. The specific operation range according to claim 3, wherein the high speed side of the high load side in the stratified charge combustion range is set, and the fuel injection control means controls the fuel injection valve when the engine is in the specific operation range. A control device for a spark ignition type direct injection engine, characterized in that fuel is divided into two equal parts and injected during a period after the middle stage of a compression stroke of a cylinder. 5. The exhaust gas recirculation passage according to any one of claims 1 to 4, which recirculates a part of exhaust gas to an intake system of the engine, an on-off valve which opens and closes the exhaust gas recirculation passage, and the engine includes at least stratified combustion. An exhaust gas recirculation control means for opening the on-off valve to recirculate the exhaust gas when in a region, the control device for a spark ignition type direct injection engine. 6. The fuel injection control means according to claim 5, when the ratio of the exhaust gas recirculation amount to the intake air amount is equal to or more than a predetermined value, the fuel injection control unit prohibits the split injection of fuel by the fuel injection valve. A control device for a spark ignition type direct injection engine, wherein the fuel injection valve is configured to inject fuel collectively in a compression stroke of a cylinder. 7. A fuel injection valve is provided at a peripheral portion of the combustion chamber while an ignition plug is provided at a ceiling portion of a combustion chamber in a cylinder of the engine, and the fuel injection is performed when the engine is in a stratified combustion region. In a control device for a spark ignition type direct injection engine, in which fuel is injected by a valve so as to face a tumble flow in a combustion chamber to stratify the air-fuel mixture around an electrode of the spark plug, the engine comprises the stratification Fuel injection for injecting fuel by the fuel injection valve in the compression stroke of the cylinder so that the injection rate on the injection start side is higher than that on the injection end side when at least on the low speed side in the high load side in the combustion region. A control device for a spark ignition type direct injection engine, which is provided with a control means. 8. The control device for a spark ignition type direct injection engine according to claim 7, wherein the fuel injection valve is configured to open and close by the operation of the piezoelectric element.