JP2001271650A - Spark ignition type direct injection engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To improve fuel consumption and output by improving combustion performance and broadening the stratification combustion range by controlling the behavior of the fuel atomization in a combustion chamber so that air-fuel mixture can be stratified properly over the broad running range of a spark ignition type direct injection engine in the stratification combustion run by the engine. SOLUTION: A tumble flow T is generated, so as to flow toward an injector between the electrode of an ignition plug 16 and the crown face of a piston 5 in a compression stroke of a cylinder 2. A recessed section 5a is provided, so as to make the tumble flow T flow along the crown face of the piston 5. The electrode of an ignition plug 16 is projected by a prescribed length from the ceiling of the combustion chamber 6. The direction of fuel injected by the injector 18 is set, so that the position P1 where the center line F of fuel atomization and the bottom face of the recessed section 5a cross each other is positioned on the exhaust side more than the position P2 where a center line z of the cylinder 2 and the bottom face of the recessed section 5a cross each other, while fuel is injected in the cylinder 2.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a spark-ignition direct-injection engine in which fuel is directly injected into a combustion chamber in a cylinder, and the mixture is ignited while being stratified around the electrodes of a spark plug. Particularly, the present invention belongs to the technical field of spray behavior control for promoting appropriate stratification of an air-fuel mixture by effectively utilizing a tumble flow in a combustion chamber.

[0002]

2. Description of the Related Art Conventionally, in this type of spark ignition type direct injection engine, a high-pressure fuel injection valve is disposed so as to face a combustion chamber in a cylinder, and a cavity having a predetermined shape is formed in a crown surface of a piston. Then, after the fuel injected from the fuel injection valve once collides with the inner wall surface or the bottom surface of the opposing cavity, the air-fuel mixture is confined in the cavity and stratified around the electrode of the ignition plug. ing.

[0003] In addition, the air-fuel mixture is collected in the cavity by utilizing the intake air flow such as swirl or tumble of the combustion chamber,
Some are transported to the vicinity of the electrode of the spark plug. For example, in the in-cylinder injection engine disclosed in Japanese Patent Application Laid-Open No. 11-161338, the fuel injection direction is reversed from the tumble flow, and the air-fuel mixture overflowing from the cavity is pushed back and confined by the tumble flow. At the same time, while promoting the vaporization of fuel droplets and mixing with air,
The mixture is transported to the vicinity of the electrode of the spark plug in the cavity.

Alternatively, as in a cylinder injection type spark ignition engine disclosed in Japanese Patent Application Laid-Open No. 11-200866,
There is also one in which fuel scattered in the cavity is transported to the spark plug side by being carried on a tumble flow. In this device, the cavity is formed substantially at the center of the crown surface of the piston, and the shape thereof is spherical so as to enhance the preservability of the tumble flow.
By widening the angle from 0 ° to 90 ° to reduce the penetration of the fuel spray, the fuel is prevented from adhering to the piston crown.

Further, SAE1999-01-0170 (Society of
Like the direct injection engine described in Automotive Engineers' technical literature), an extremely strong tumble flow is generated in the combustion chamber inside the cylinder, and the tumble flow ignites the fuel spray from the fuel injector along the injection direction. Some are transported near the electrode of the plug. In this case, as shown in FIG. 46 (a), fuel is injected by a fuel injection valve so as to spread relatively large, and as shown in FIG. 46 (b), the fuel spray is close to the combustion chamber ceiling. The upper one is carried on the tumble flow and transported to the spark plug side, while the lower one close to the piston crown collides with the tumble flow to reduce its momentum. Then, as shown in FIG. 3 (c), the entire fuel spray is put on a tumble flow and transported to the spark plug side along the ceiling of the combustion chamber.

[0006]

However, it is difficult to say that any of the above-mentioned conventional engines properly stratifies the air-fuel mixture over a wide operating range in which the load state and the rotation speed are different. When viewed over the entire operation range, there is considerable room for improving the effects such as fuel efficiency improvement by the stratified combustion operation. That is, as in the first conventional example (Japanese Patent Laid-Open No. 11-161338), it is assumed that the air-fuel mixture is confined in the cavity.
The operating area where the mixture can be stratified properly is severely restricted by the dimensions and shape of the cavity, and the area where the engine can be stratified is actually limited to a narrow area on the low-load and low-speed side, which also improves the fuel efficiency. It cannot be too large.

In addition, in such a direct injection engine, generally, the inner wall surface of the cavity where the fuel spray from the fuel injection valve collides is located near the center line of the cylinder.
Since it is inevitable that the growth of the flame nucleus in the early stage of combustion is reduced by the inner wall surface, and the propagating property of the flame is reduced, the fact is that the combustibility is reduced by this. In addition, in this case, as described above, the fuel spray is caused to collide with the inner wall surface and the bottom surface of the cavity, so that the amount of fuel adhering to the wall surface and the like increases, which leads to deterioration of fuel efficiency and exhaust gas. There is also a disadvantage that unburned hydrocarbons (HC) increase.

For example, FIG. 47 shows that a plurality of pistons having different shapes of cavities are prepared, and a fuel efficiency improvement rate and an output improvement rate of an engine by direct injection are respectively experimentally obtained.
The results are shown in comparison. According to FIG.
There is a so-called trade-off relationship between the improvement of fuel efficiency and the improvement of output. In the former case in which a deep dish-shaped cavity is provided (point A), the fuel spray is appropriately trapped and stratified. Therefore, the effect of improving fuel efficiency in a low-load and low-speed range is high, but the combustibility is reduced particularly on a high-speed side, and the effect of improving output is low.

On the other hand, with a so-called flat piston having only a concave crown surface (point C), although the output improvement effect on the high rotation side is high, it is difficult to appropriately stratify the air-fuel mixture in a low load region. However, the effect of improving fuel efficiency must be reduced. And as an intermediate between them,
When the inner wall surface of the cavity facing the fuel injection valve is greatly inclined (point B), a significant improvement in both fuel consumption and output cannot be expected.

Next, a second conventional example (Japanese Patent Laid-Open No. 11-2008)
Looking at No. 66), it is considered that, even in this case, when the size of the cavity is small, the same problem as in the former conventional example occurs. On the other hand, when the size of the cavity is large, the air-fuel mixture is similar to the flat piston. Proper stratification becomes difficult. In other words, in this device, the fuel scattered in the cavity is put on the tumble flow and is transported to the ignition plug side. Even if the fuel can be collected, the fuel moves on the tumble flow and passes near the electrode of the spark plug, so that the period during which the fuel mixture can be ignited by the spark plug is extremely short. It cannot be said that appropriate stratification has been achieved.

It is considered that such a problem becomes more prominent in the third conventional example (SAE1999-01-0170). That is, in this conventional example, the fuel injection direction is twisted by an extremely strong tumble flow, and the entire fuel spray is transported by the tumble flow. Therefore, the speed of the fuel spray moving on such a strong tumble flow is This is because the air-fuel mixture naturally increases, and the air-fuel mixture passes near the electrode of the spark plug in a very short time.

The present invention has been made in view of the above points, and a main object of the present invention is to perform a stratified charge combustion operation using a spark ignition type direct injection engine over a wide operating range of the engine. In order to appropriately stratify the air-fuel mixture, the fuel spray behavior in the combustion chamber is controlled, and the fuel efficiency and output are improved by improving the combustibility and expanding the stratified combustion region.

[0013]

In order to achieve the above object, the present invention focuses on a tumble flow in a compression stroke of a cylinder, and forms a recess along the crown surface of the piston along the tumble flow. The fuel injection direction for the tumble flow and the position of the electrode of the ignition plug are optimally set, and the behavior of the fuel spray is controlled by the tumble flow to appropriately stratify the air-fuel mixture around the electrode of the ignition plug. I did it.

Specifically, in the invention of claim 1, a spark plug is provided on a ceiling of a combustion chamber facing a crown of a piston in a cylinder so as to protrude from near the center of the ceiling toward the combustion chamber. A fuel injection valve is provided so as to inject fuel from the peripheral portion of the ceiling toward the center of the combustion chamber, and the fuel injected from the fuel injection valve is used for the ignition plug during stratified combustion operation. It is assumed that a spark-ignition direct-injection engine stratified around the electrodes. And
Tumble flow generating means capable of generating a tumble flow that flows between the electrode of the spark plug and the crown of the piston during the compression stroke of the cylinder toward the fuel injection valve; At the predetermined time after the fuel injection start timing of the fuel injection valve during the compression stroke of the cylinder and before the ignition timing of the cylinder, the piston is provided with a recess. It is located at an intermediate portion between the tumble forward flow flowing along the concave portion of the crown surface and the tumble forward flow flowing along the combustion chamber ceiling. The direction in which the fuel is injected by the fuel injection valve is such that, at the fuel injection timing of the cylinder during the stratified charge combustion operation, the position where the center line of the fuel spray intersects the bottom surface of the recess is defined by the cylinder center line and the bottom surface of the recess The fuel injection valve is set so as to be farther than the intersection position.

With the above configuration, during stratified charge combustion operation of the engine, of the tumble flow generated by the tumble flow generating means, a positive tumble flow flowing along the concave portion of the piston crown surface in the compression stroke of the cylinder is supplied to the fuel injection valve. With the tumble forward flowing along the ceiling of the combustion chamber, the cylinder is maintained without collapse until the latter half of the compression stroke of the cylinder. When fuel is injected by the fuel injection valve in this state, the fuel spray collides with the tumble positive flow to promote atomization and dispersion of the fuel or mixing with the surrounding air, and It moves against the tumble positive flow along the concave portion of the piston crown surface, and becomes a flammable mixture of an appropriate concentration in the vicinity of the electrode of the spark plug and stays there.

Here, since the center line of the fuel spray from the fuel injection valve is relatively far and intersects the crown of the piston, the fuel spray does not adhere much to the crown of the piston. Instead, it effectively collides with a strong portion of the tumble flow that flows along the concave portion, whereby the behavior of the fuel spray can be accurately controlled by the tumble flow. Further, since the electrode of the spark plug protrudes greatly into the combustion chamber and is located near the center of the vortex of the tumble flow, the vicinity of the electrode of the spark plug is in a state where the air-fuel mixture is easily retained.

Therefore, according to the present invention, the behavior of the fuel spray can be controlled by the tumble flow as described above, and the fuel spray can be appropriately stratified around the electrode of the ignition plug.
The inner wall surface of the cavity for colliding the fuel spray as in the conventional example becomes unnecessary, and the flammability is increased by this, so that the fuel efficiency can be reduced. In addition, since there is no restriction on the dimensions and shape of the cavity as in the conventional example, a favorable stratified combustion state can be achieved up to a higher rotation speed range, so that a significant improvement in fuel efficiency can be achieved in the entire operating range of the engine. become. Further, as described above, the combustible air-fuel mixture can be retained near the electrode of the ignition plug, and the time in which the air-fuel mixture can stably ignite is sufficiently long. As a result, the fuel efficiency and the output can be improved.

According to the second aspect of the present invention, the electrode of the spark plug is located closer to the piston crown than the injection hole of the fuel injection valve when viewed from a direction perpendicular to the cylinder center line. Thereby, the positions of the electrodes of the ignition plug are embodied, and the operation and effect of the first aspect of the invention can be sufficiently obtained.

According to the third aspect of the present invention, the electrode of the spark plug projects from the ceiling of the combustion chamber along the cylinder center line, and the cylinder extends from the ceiling of the combustion chamber to the piston crown surface at the compression top dead center of the cylinder. When the distance on the center line is d, the distance e from the combustion chamber ceiling to the electrode of the spark plug is
It is set to a value that satisfies the relationship e ≧ 0.4d. Thereby, the position of the electrode of the spark plug is embodied, and
The effect of the invention of the invention is sufficiently obtained.

According to the fourth aspect of the present invention, the direction of fuel injection by the fuel injection valve is set so that the center line of the fuel spray passes between the electrode of the spark plug and the recess on the piston crown. By doing so, the fuel spray from the fuel injection valve reverses the tumble positive flow guided by the concave portion of the piston crown surface, and reliably stays near the electrode of the spark plug. The effect of the invention of the invention is sufficiently obtained.

According to the fifth aspect of the present invention, the fuel injection direction of the fuel injection valve is set at an inclination angle of about 25 ° to about 40 ° with respect to an imaginary plane in which the center line of the fuel spray is perpendicular to the cylinder center line. It shall be set as follows.

In this way, by setting the center line of the fuel spray to approximately 40 ° or less with respect to an imaginary plane orthogonal to the cylinder center line, the tumble positive flow flowing along the concave portion of the piston crown surface can be prevented. The collision can be made such that the center of the fuel spray is substantially directly opposed. Further, by setting the center line of the fuel spray to be approximately 25 ° or more with respect to the virtual plane, the fuel spray is less affected by the tumble forward flow flowing along the ceiling of the combustion chamber. Therefore, the behavior of the fuel spray can be controlled with higher accuracy by the tumble positive flow.

According to the sixth aspect of the present invention, the shape of the fuel spray during the compression stroke of the cylinder is larger when viewed along the cylinder center line than when viewed along the axis of the crankshaft. It is assumed that the electrode has an expanded flat shape and that the electrode of the ignition plug is located outside the geometrical fuel spray area from the fuel injection valve in the compression stroke of the cylinder. Note that the geometric spray area is an area of fuel spray droplets when there is no intake air flow in the combustion chamber.

In this configuration, the spread angle of the fuel spray in the direction in which the cylinder center line extends is relatively small when viewed along the axis of the crankshaft. Can collide effectively,
In addition, the influence of the tumble forward flow can be reduced. Thus, the accuracy of the behavior control of the fuel spray by the positive tumble flow can be improved. Also, by positioning the spark plug electrode outside the geometric fuel spray area,
By preventing fuel droplets from adhering to the electrodes, it is possible to prevent smoldering of the spark plug.

According to the seventh aspect of the present invention, in the fuel injection valve, the penetration force of the fuel spray during the compression stroke of the cylinder is larger on the side closer to the crown of the piston than the opposite side with respect to the center line of the fuel spray. .

In this way, if the side of the fuel spray that is closer to the piston crown surface is made to substantially directly collide with the tumble positive flow, the spread angle of the fuel spray itself is relatively large, even if it is relatively large. Can be effectively collided with the tumble positive flow, and the accuracy of the fuel spray behavior control by the tumble positive flow can be improved. On the other hand, the side of the fuel spray near the ceiling of the combustion chamber has a small spray penetration force, so that the amount of air-fuel mixture which is diffused by the fuel spray flowing by the tumble forward flow is reduced, and the fuel efficiency is improved due to the improvement of the degree of stratification. Is improved.

In the fuel injection valve of the present invention, the fuel spray penetrating force in the compression stroke of the cylinder is larger on the side closer to the combustion chamber ceiling with respect to the center line of the fuel spray than on the opposite side. In addition, it is assumed that the electrode of the ignition plug is located outside the geometric fuel spray area from the fuel injection valve in the compression stroke of the cylinder.

In this way, if the side of the fuel spray near the ceiling of the combustion chamber is made to substantially directly collide with the tumble positive flow, the spread angle of the fuel spray itself is relatively large, Many fuels can effectively collide with the tumble positive flow. On the other hand, the side of the fuel spray close to the piston crown has a small spray penetration force, so that adhesion of fuel to the piston crown can be suppressed. Therefore, the accuracy of fuel spray behavior control by the tumble positive flow can be improved.
Further, by positioning the electrode of the spark plug outside the geometric fuel spray area, it is possible to prevent fuel droplets from adhering to the electrode and prevent smoldering of the spark plug.

According to the ninth aspect of the present invention, the recess in the crown surface of the piston is elongated in the direction in which the center line of the fuel spray extends when viewed along the cylinder center line, and at least substantially in the length direction of the recess. The width of the central portion is made wider than the width of the end portion on the side close to the fuel injection valve, and the side wall surfaces on both sides in the width direction are protruded from the bottom surface to a predetermined height or more at the substantially central portion in the longitudinal direction of the concave portion. I do.

By extending the shape of the concave portion in the direction in which the center line of the fuel spray extends, the direction of the tumble flow flowing along the concave portion is directed toward the fuel spray. In addition, the tumble flow can stably collide with the fuel spray. In addition, by making the height of the side wall surface of the concave portion equal to or more than a predetermined value, the diffusion of the fuel spray can be directly suppressed by the both side wall surfaces, and the side surface of the concave portion is directed from the outside to the inside of the concave portion. A squish flow is generated, and the squish flow also suppresses fuel diffusion.

In the tenth aspect of the present invention, the width of the end farthest from the fuel injector among the two ends in the longitudinal direction of the recess in the ninth aspect is set to the width of the substantially central part in the longitudinal direction of the recess. And it is almost the same. This effectively suppresses the attenuation of the tumble flow flowing from the ceiling of the combustion chamber to the crown of the piston along the inner peripheral surface of the cylinder, so that the tumble flow is smoothly introduced into the recess. As a result, the flow velocity of the tumble flow for controlling the behavior of the fuel spray can be sufficiently ensured.

According to the eleventh aspect of the present invention, of the end portions on both sides in the longitudinal direction of the concave portion according to the ninth aspect, the width of the end portion farther from the fuel injection valve is set to be wider than the substantially central portion in the longitudinal direction of the concave portion. Be wide. Thus, the tumble flow can be more smoothly introduced into the concave portion as compared with the invention of the tenth aspect, so that the flow velocity of the tumble flow can be more easily secured.

In the twelfth aspect of the present invention, a squish area is formed at least on the outer peripheral side of the crown surface of the piston on the exhaust side so as to extend along the ceiling of the opposed combustion chamber. As a result, the squish flowing from the gap (squish area) between the squish area of the crown of the piston and the ceiling of the combustion chamber toward the center of the combustion chamber with the rise of the piston near the ignition timing of the cylinder. A flow is generated, and the diffusion of the air-fuel mixture at the center of the combustion chamber is suppressed by the squish flow, so that the air-fuel mixture can stay longer near the electrode of the spark plug.

According to the thirteenth aspect of the present invention, the fuel injection valve is located near the intake valve of the intake and exhaust valves, and the concave portion of the piston crown surface is viewed along the cylinder center line. The shape is long in the direction in which the center line of the fuel spray extends, and the volume of the concave portion is smaller on the side closer to the fuel injection valve than the cylinder center line on the side opposite to the fuel injection valve. As a result, the combustion chamber volume on the intake valve side, where the temperature is relatively low and flame propagation tends to be slower, is made smaller than that on the exhaust valve side, so that flame propagation is uniformed and abnormalities on the intake side. Combustion can be prevented and knocking can be suppressed.

[0035]

Embodiments of the present invention will be described below with reference to the drawings.

(Embodiment 1) FIG. 2 shows an overall configuration of a spark ignition type direct injection engine 1 according to Embodiment 1 of the present invention. The 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 disposed on the cylinder block 3. A piston 5 is inserted into each of the cylinders 2 so as to be able to reciprocate up and down in the figure. The combustion chamber 6 is partitioned in the cylinder 2 between the two. On the other hand, a crankshaft 7 is rotatably supported in the cylinder block 3 below the piston 5, and the crankshaft 7 and the piston 5 are connected by a connecting rod 8. At one end of the crankshaft 7, an electromagnetic crank angle sensor 9 for detecting the rotation angle is provided.

As shown in the enlarged view of FIG. 3, two inclined surfaces extending from a substantially central portion to near the lower end surface of the cylinder head 4 are formed on the ceiling of each of the cylinders 2. Is a so-called pent-roof type combustion chamber 6 having a shape like a roof that is laid over each other. In other words, the height dimension of the combustion chamber 6, that is, the distance parallel to the cylinder center line z from the piston crown surface to the combustion chamber ceiling portion is at least on the cylinder center line z more than the peripheral portions on the intake and exhaust sides. It is getting bigger. Two intake ports 10 and two exhaust ports 11 are respectively opened on the two inclined surfaces, and intake and exhaust valves 12, 12, 13, 13 are arranged at the opening ends of the respective ports. The two intake ports 10, 10 each extend linearly obliquely upward from the combustion chamber 6 and open independently on one side (right side in FIG. 2) of the engine 1. The two exhaust ports 11, 11 merge into one at the middle and extend substantially horizontally, and open to the other side of the engine 1 (the left side in FIG. 2).

The intake valve 12 and the exhaust valve 13 are opened by being pressed in the valve axis direction by two camshafts (not shown) supported in the cylinder head 4.
The two camshafts are each rotated by the timing belt in synchronization with the crankshaft 7, so that the intake valve 12
The exhaust valve 13 is opened and closed at a predetermined timing for each cylinder 2. The two camshafts are respectively provided with known variable valve mechanisms 14, 14 for continuously changing the rotational phase with respect to the crankshaft 7 within a predetermined angle range. 14, the opening / closing operation timing of the intake valve 12 and the exhaust valve 13 is independently changed.

As shown in FIG. 3, an ignition plug 16 is disposed above the combustion chamber 6 so as to be surrounded by the four intake and exhaust valves 12 and 13. The electrode at the tip of the spark plug 16 is located at a position protruding from the ceiling of the combustion chamber 6 by a predetermined distance, while an ignition circuit 17 (only shown in FIG. 2) is connected to the base end of the spark plug 16. The ignition plug 16 is energized at a predetermined ignition timing for each cylinder 2. Further, the combustion chamber 6
The crown surface of the piston 5 which is the bottom of the combustion chamber 6 is shaped so as to be along the ceiling of the opposed combustion chamber 6, and a lemon-shaped recess 5a is provided at the center thereof.

An injector (fuel injection valve) 18 is disposed on the periphery of the combustion chamber 6 so as to be sandwiched below the two intake ports 10 and 10. The injector 18 is a well-known swirl injector that injects fuel as a swirling flow from an injection hole at a tip end thereof and injects the fuel in a hollow cone shape along the direction in which the axis of the injector 18 extends. In the swirl injector 18, when the injection pressure of the fuel is increased, the penetration force of the fuel spray is increased in accordance with the increase in the pressure. In addition, as shown in FIG. 4, as the spread angle of the spray increases, the penetration force of the spray tends to decrease. Conversely, when the spray angle decreases, the penetration force increases. In addition, the injector 18
Is provided with a variable mechanism (spray angle variable mechanism) for varying the strength of the swirl flow component of the fuel. By operating the variable mechanism, the swirl flow component of the fuel is adjusted to change the spread angle of the fuel spray. It may be.

A fuel distribution pipe 19 common to all the cylinders 2, 2,... It is supposed to. More specifically, the fuel supply system 20 is configured, for example, as shown in FIG. 23, a low-pressure regulator 24, a fuel filter 25, a high-pressure fuel pump 26, and a high-pressure regulator 27 are arranged in this order.

The fuel sucked from the fuel tank 21 by the low-pressure fuel pump 23 is supplied to the low-pressure regulator 2.
The pressure is adjusted by the fuel filter 4, filtered by the fuel filter 25, and fed to the high-pressure fuel pump 26. This high pressure fuel pump 26
And the high-pressure regulator 27 are connected to the return passage 29, respectively.
Is connected to the fuel tank 21 side, and while returning a part of the fuel pressurized by the high pressure fuel pump 26 to the fuel tank 21 side by the return passage 29 while adjusting the flow rate by the high pressure regulator 27, the fuel is distributed to the fuel distribution pipe 19. Properly adjust the pressure state of the supplied fuel (for example, from about 3 MPa to
Approx. 13 MPa, preferably 4M during stratified combustion operation
(To about 7 MPa).
The return passage 29 is provided with a low-pressure regulator 28 for adjusting the pressure state of the fuel returned to the fuel tank 21. In the fuel supply system 20, the high-pressure fuel pump 26 and the high-pressure regulator 27 constitute injection pressure adjusting means for adjusting the injection pressure of fuel by the injector 18.

The structure of the fuel supply system 20 is the same as that shown in FIG.
The high pressure regulator 27 is not limited to the one shown in (a) but may be omitted, for example, as in a fuel supply system 20 'shown in FIG. In this case, the electric high-pressure pump 29 capable of changing the fuel discharge amount over a wide range.
From the electric high pressure pump 29 to the fuel distribution pipe 19
The pressure state of the fuel can be controlled by variably adjusting the amount of fuel discharged to the fuel cell.

As shown in FIG. 2, an intake passage 30 is connected to one side of the engine 1 so as to communicate with the intake ports 10 of each cylinder 2. This intake passage 3
Numeral 0 denotes the supply of intake air filtered by an air cleaner (not shown) to the combustion chamber 6 of the engine 1. Wire type air flow sensor 31
, An electric throttle valve 32 for restricting the intake passage 26, and a surge tank 33 are provided. The electric throttle valve 32 is not mechanically connected to an accelerator pedal (not shown), and is driven by an electric drive motor (not shown) to open and close.

Further, the intake passage 30 downstream of the surge tank 33 is an independent passage branching 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. As shown in FIG. 3, an intake flow control valve 34 for adjusting the flow rate of the tumble flow in the combustion chamber 6 is disposed upstream of both of the two intake ports 10, 10.
For example, it is opened and closed by a stepping motor 35 (shown only in FIG. 3). Each of the intake flow control valves 34, 34 is formed by cutting out a part of a circular butterfly valve. In this embodiment, a portion below the valve shaft 34a is cut out. Then, when the intake flow control valve 34 is closed, the intake air flows downstream only from the notch portion, and generates a strong tumble flow in the combustion chamber 6. On the other hand, as the intake flow control valve 34 is opened, the intake air flows from other than the notched portion, and the strength of the tumble flow is gradually reduced.

The intake ports 10 and 10 constitute a tumble flow generating means for generating a tumble flow in the combustion chamber 6. The tumble flow is generated between the electrode of the ignition plug 16 and the piston 5 during the compression stroke of the cylinder 2 as described later. It flows between the crown surface and the injector 18. The intake flow control valve 34 and the stepping motor 35
Thus, the tumble flow rate varying means for varying the flow rate of the tumble flow is configured. The shapes of the intake port 10 and the intake flow control valve 34 are not limited to those described above.
For example, the intake port may be a so-called common port that joins one at the upstream side. In this case, the intake flow control valve may have a shape corresponding to the cross-sectional shape of the common port as a base, and a shape in which a part of the butterfly valve is cut out in the same manner as described above.

On the other hand, the combustion chamber 6
An exhaust passage 36 for discharging burned gas (exhaust gas) from the exhaust gas is connected. The upstream end of the exhaust passage 36 is branched for each cylinder 2 and is connected to the exhaust port 11.
7, a linear O 2 sensor 38 for detecting the oxygen concentration in the exhaust gas is provided at the gathering portion of the exhaust manifold 37. The linear O2 sensor 38 is used for detecting the air-fuel ratio based on the oxygen concentration in the exhaust gas, so that a linear output with respect to the oxygen concentration can be obtained in a predetermined air-fuel ratio range including the stoichiometric air-fuel ratio. ing.

An exhaust pipe 39 is connected to an upstream end of an exhaust pipe 39, and a downstream end of the exhaust pipe 39 is connected to a catalyst 40 for purifying exhaust gas.
Is connected. The catalyst 40 absorbs NOx in an atmosphere having a high oxygen concentration in the exhaust gas, and releases NOx absorbed by reduction of the oxygen concentration and reduces and purifies NOx.
It is of the x-absorption reduction type, and exhibits a high exhaust gas purification performance similar to that of a so-called three-way catalyst particularly near the stoichiometric air-fuel ratio. Further, a known lambda O2 sensor 41 whose output is reversed in a stepwise manner at the boundary of the stoichiometric air-fuel ratio is disposed downstream of the catalyst 40 in order to determine the deterioration state of the catalyst 40. In addition, in addition to the NOx absorption reduction type catalyst 40, a three-way catalyst may be arranged in series with the NOx absorption reduction catalyst.

Further, an upstream end of an EGR passage 43 for recirculating a part of the exhaust gas flowing through the exhaust passage 36 to the intake passage 30 is branched and connected upstream of the exhaust pipe 39. The downstream end of the EGR passage 43 is connected to the intake passage 30 between the throttle valve 32 and the surge tank 33, and an electric EGR valve 44 whose opening can be adjusted is disposed near the intake passage 30. The amount of exhaust gas recirculated by the passage 43 can be adjusted.

The variable valve mechanism 14, the ignition circuit 17 of the ignition plug 16, the injector 18, the high-pressure regulator 27 of the fuel supply system 20, the electric throttle valve 32, the intake flow control valve 34, the electric EGR valve 44, etc. In each case, an engine control unit 50 (hereinafter referred to as ECU)
The operation is controlled by the On the other hand, the ECU 50 receives at least each output signal from the crank angle sensor 9, the air flow sensor 31, the linear O2 sensor 38, the lambda O2 sensor 41, and the like. ) And an output signal from a rotation speed sensor 52 that detects the rotation speed of the engine 1 (the rotation speed of the crankshaft 7).

The ECU 50 adjusts the opening / closing operation timing of the intake / exhaust valves 12, 13, the fuel injection amount by the injector 18, the injection timing and injection pressure, and the throttle valve 32 based on signals input from the respective sensors. The amount of intake air, the strength of the tumble flow adjusted by the intake flow control valve 34, the recirculation ratio of the exhaust adjusted by the EGR valve 44, and the like are controlled in accordance with the operating state of the engine 1, respectively.

Specifically, for example, as shown in FIG. 6, when the engine 1 is in the warm state, the set operation region (a) on the low-load and low-rotation side is a stratified combustion region. At a predetermined time in the compression stroke (for example, approximately four times before compression top dead center (BTDC) during stratified combustion operation).
The fuel injection is performed in the range of 0 ° CA to approximately 140 ° CA), and a stratified charge combustion mode is set in which the mixture is burned in a state where the mixture is unevenly distributed in the vicinity of the ignition plug 16. In this stratified combustion state, the throttle valve 3 is used to reduce the intake loss of the engine 1.
The opening degree of the fuel cell 2 is made relatively large. At this time, the average air-fuel ratio of the combustion chamber 6 becomes leaner than the stoichiometric air-fuel ratio (for example, A / F> 25).

On the other hand, a region other than the stratified fuel region (b)
Is a homogeneous combustion region, in which fuel is injected by the injector 17 in the intake stroke of the cylinder 2 and is sufficiently mixed with the intake air to form a uniform mixture in the combustion chamber 6 and then burn. . In this uniform combustion state, the air-fuel ratio of the air-fuel mixture is substantially equal to the stoichiometric air-fuel ratio (A / F) in most of the operating range.
The fuel injection amount, throttle opening, and the like are controlled so as to achieve (≒ 14.7). Particularly, in the full load operation state, the air-fuel ratio is made richer than the stoichiometric air-fuel ratio (for example, A / F = 13).
Control) to obtain a large output corresponding to a high load.

Further, when the engine 1 is warm, the EGR valve 44 is opened in a region shown by hatching in FIG. . At this time, the opening degree of the EGR valve 44 is adjusted according to the load state and the rotation speed of the engine 1 so that the exhaust gas recirculation ratio (hereinafter, also referred to as the EGR rate) becomes smaller at least on the higher load side. With this,
The generation of NOx can be suppressed by the recirculation of exhaust gas without impairing the combustion stability of the engine 1. In addition, when the engine is cold, the highest priority is given to ensuring combustion stability, and a uniform combustion state is set in all operating regions of the engine 1 and the EGR valve 44 is fully closed.

As the EGR rate, for example, the ratio of the amount of exhaust gas recirculated to the intake passage 30 by the EGR passage 43 to the amount of fresh air may be used. Here, fresh air is outside air obtained by removing the recirculated exhaust gas, fuel gas, and the like from the gas sucked into the cylinder 2.

The feature of the present invention is that the engine 1
When the is operated in a stratified combustion state, the tumble flow in the combustion chamber 6 is utilized to the utmost, the behavior of the fuel spray is controlled by the tumble flow, and the stratified mixture is appropriately stratified. is there.

That is, when the engine 1 is in the stratified combustion region (A), the tumble flow generated in the intake stroke of each cylinder 2 is held until the latter stage of the compression stroke of the cylinder 2 and the tumble flow is substantially positive. Fuel is injected so as to collide from the opposite direction. In this way, the fuel spray moves toward the spark plug 16 while being gradually decelerated by the tumble flow, and during that time, the vaporization and atomization of the fuel droplets and the mixing with the air are promoted, and the fuel spray is moved at the ignition timing of the cylinder 2. As shown in FIG.
It stays near the electrode of the ignition plug 16.

In other words, the penetration force of the fuel spray by the injector 18 is adjusted so as to correspond to the flow velocity of the tumble flow, and the injector 18 is operated at a predetermined timing calculated backward from the ignition timing of the cylinder 2 to discharge the fuel. I am trying to inject it. Such operation control of the injector 18 is performed by the ECU 50 in accordance with a predetermined control program as described above.
0 adjusts the penetration force of the fuel spray by the injector 18 so as to correspond to the flow rate of the tumble flow, and the fuel spray reverses the tumble flow to become a combustible mixture at the time of ignition by the spark plug 16. This corresponds to a fuel injection control means for injecting fuel by the injector 18 in association with the ignition timing of the cylinder 2 so as to stay near the electrode of the ignition plug 16.

Specifically, in this embodiment, the shape of the crown surface of the piston 5, the direction of fuel spray by the injector 18, the spray spread angle, and the position of the electrode of the spark plug 16 are optimally set in relation to each other. Above,
By adjusting the penetration force of the fuel spray by the injector 18 according to the flow velocity of the opposed tumble flow, appropriate stratification of the air-fuel mixture as described above is realized. Hereinafter, each of the above-mentioned features will be specifically described.

(Shape of Piston Crown Surface) First, as shown in FIGS. 7 and 8 and FIGS. 1 and 3, the crown surface of the piston 5 was viewed along the cylinder center line z. sometimes,
A long concave portion 5a is formed in the fuel injection direction of the injector 18 (the direction in which the center line of the fuel spray extends). Further, as shown in FIG. 8, the widthwise dimension of the concave portion 5a is set to include the fuel spray from the injector 18 as a whole (indicated by diagonal lines in FIG. 8).

More specifically, the width w of the concave portion 5a is the widest at a substantially central portion in the length direction, and gradually narrows from both ends toward both ends. The bottom surface of the concave portion 5a has a substantially arc shape when viewed from a direction orthogonal to the cylinder center line z as shown in FIG. 7B, and the deepest portion corresponds to the cylinder center line z. That is, it is located substantially at the center of the cross section of the cylinder 2. Further, as shown in FIG. 3 (c), the side walls on both sides in the width direction of the recess 5a gradually rise up from the bottom and extend upward, and the height of the side wall is
The height is set to be the highest in the substantially central portion in the length direction of “a”, and is equal to or higher than a predetermined value so as to suppress the diffusion of the fuel spray by the injector 18.

In addition, the piston 5 excluding the recess 5a
The outer peripheral portion of the crown surface is formed as a squish area having a shape substantially parallel to the inclined surface of the ceiling of the opposed combustion chamber 6, and is provided for a predetermined period (for example, BT) before the compression top dead center of the cylinder 2.
At DC 40 ° CA to TDC), a squish area is a gap sandwiched between the outer peripheral portion of the piston 5 crown surface and the ceiling of the combustion chamber 6.

By providing the recess 5a on the crown surface of the piston 5, the flow of the tumble flow in the combustion chamber 6 becomes smooth, and the tumble flow T generated during the intake stroke of the cylinder 2 (see FIG. 32). ) Is surely held without being attenuated until late in the compression stroke of the cylinder 2. When fuel is injected by the injector 18 as shown in FIG. 8, the tumble flows T, T (tumble flow)
Flows along the side wall surface of the concave portion 5a so as to bind the fuel spray to the inside, thereby suppressing the diffusion of the fuel spray and stably colliding with the fuel spray to control the behavior thereof.
To stay in the center. Further, although not shown, in the vicinity of the compression top dead center of the cylinder 2 (for example, after BTDC 20 ° CA), the cylinder 2 is moved from the squish area on the outer peripheral side of the crown surface of the piston 5 toward the cylinder center line z as described above. A strong flowing squish flow is generated, which effectively suppresses the diffusion of the mixture.

(Arrangement of Injector) Second, the arrangement of the injector 18 will be described. First, as shown in FIG. 9, the injector 18 is disposed such that its axis (the center line F of the fuel spray) forms an inclination angle δ of approximately 30 ° with respect to a virtual plane orthogonal to the cylinder center line z. I have. In other words, the center line F of the fuel spray by the injector 18
Is the fuel injection timing of the cylinder 2 during the stratified combustion operation (BTDC 40
(° CA to 140 ° CA), a position (point P) that passes between the electrode of the spark plug 16 and the concave portion 5a of the crown surface of the piston 5 and intersects the spray center line F and the bottom surface of the concave portion 5a.
1) is set farther from the injector 18 than a position (point P2) where the cylinder center line z intersects the bottom surface of the concave portion 5a. Although the spread angle θ of the fuel spray by the injector 18 generally changes depending on the pressure state of the combustion chamber 6, in this embodiment, the spray spread angle θ in the compression stroke of the cylinder 2 is represented by θ = approximately 20 °
It is set to fall within a range of about 60 °.

Here, the definition of fuel spray in this specification will be described with reference to FIG. As shown in FIG. 3A, two points B and C at which a virtual plane passing through the spray center line F and the contour of the fuel spray intersect are determined at a position 20 mm downstream from the injection hole point A of the injector 18. , @BAC
Defines the spray spread angle θ (θ = ∠BAC).
Further, as shown in FIG. 3B, on the virtual plane passing through the spray center line F, the top of the fuel spray at the tip of the main spray (the area of the fuel droplets) excluding the so-called advance spray (initial spray). The tips are point B and point C, respectively.
The distance along the spray center line F from the injection hole portion A to the point B is defined as L1, the distance from the point A to the point C is defined as L2, and the average distance between them defines the spray penetration force L ( L = (L1 + L2) / 2).

As an actual measuring method of the spray spread angle θ and the spray penetration force L, for example, a laser sheet method may be used. That is, first, as a fluid to be injected by the injector, a dry solvent corresponding to fuel properties is used, and the pressure of this sample is set to a predetermined value (for example, 7 MPa) within the range of the fuel pressure actually used at normal temperature.
Set to. As the atmospheric pressure, the inside of a pressure vessel having a laser passage window capable of photographing the spray and a measurement window is pressurized to, for example, 0.25 MPa.

At room temperature, the injector 1 is controlled so that the injection amount per pulse is 9 mm ³ / stroke.
A driving pulse signal having a predetermined pulse width is input to 8 and fuel is injected. At this time, the fuel spray is irradiated with a laser sheet light having a thickness of 5 mm so as to pass through the spray center line, and a spray image is taken with a high-speed camera from a direction perpendicular to the laser sheet light plane. I do. Then, based on the photographing screen 1.56 milliseconds after the input timing of the drive pulse signal, the divergence angle θ and the penetration force L of the spray are determined according to the above definition.

Note that the outline of the spray in the photographed image is the outline of the area of the droplet-shaped sample particles, and the area of the sample particles becomes brighter by the laser sheet light. The outline of the spray is determined from the part where it is located.

Then, the injector 1 as described above
8, that is, by setting the inclination angle δ of the spray center line F and the spray divergence angle θ to the optimum values as described above, the engine 1 of this embodiment has the piston at the fuel injection time shown in FIG. The fuel spray can be made to effectively collide with the tumble flow T (positive tumble flow) flowing along the concave portion 5a of the five crown surface so that the center portion thereof (a predetermined range surrounding the spray center line F) substantially faces. This makes it possible to move the fuel spray along the concave portion 5a while attenuating the moving speed of the fuel spray by the tumble flow T, so that the fuel spray can reliably stay near the electrode of the ignition plug 16.

(Setting of Fuel Spray Center Line) The setting of the inclination angle δ of the fuel spray center line F by the injector 18 will be described more specifically with reference to FIGS. For example, when the inclination angle δ of the spray center line F is set to be smaller than approximately 25 ° (shown as a dashed line F1 in FIG. 12), most of the fuel spray from the injector 18 flows along the ceiling of the combustion chamber 6. By the flow (tumble forward flow Ts), the fuel is transported to the ignition plug 16 side, and the remaining fuel collides with the tumble flow (positive tumble flow Tm) flowing along the concave portion 5a of the piston 5 crown surface,
Thereafter, the vehicle moves on the ceiling side of the combustion chamber 6 on the tumble positive flow Tm. As a result, as shown by the dashed line in FIG. 13, the local air-fuel ratio near the electrode of the ignition plug 16 temporarily enters the combustible range when the air-fuel mixture transported by the tumble flows Ts, Tm arrives, but immediately becomes excessive. It becomes rich and returns to a state leaner than the flammable range in a very short time. That is, the spark plug 16
The local air-fuel ratio near the electrode fluctuates drastically.

On the other hand, as shown by the two-dot chain line F2 in FIG.
When the angle is larger than °, most of the fuel spray from the injector 18 collides with the crown of the piston 5 and adheres to the piston 5 as shown by the broken line in FIG.
As shown by the two-dot chain line, the local air-fuel ratio in the vicinity of the electrode of the ignition plug 16 does not fall within the flammable range, and the ignition becomes impossible. For example, there is a relationship as shown in FIG. 14 between the amount of fuel attached to the crown surface of the piston 5 and the inclination angle δ of the spray center line F, that is, the installation angle of the injector 18. The installation angle of the injector 18 is 40 °
It can be seen that the fuel adhesion amount sharply increases when the pressure exceeds.

That is, in order to stably maintain the concentration state of the air-fuel mixture in the combustible range, it is preferable that at least the inclination angle δ of the spray center line F be a value in the range of approximately 25 ° to approximately 40 °. It can be said.

For example, as in this embodiment, the inclination angle δ of the fuel spray center line F by the injector 18 is approximately 30 °.
In this case, the fuel spray from the injector 18 comes to substantially collide with the tumble flow T and effectively collide with each other, so that the movement speed of the fuel spray can be accurately adjusted by the tumble flow T. Thereby, the air-fuel mixture can be stably retained near the electrode of the ignition plug 16. That is, as shown by the solid line in FIG. 13, in this embodiment, the local air-fuel ratio near the electrode of the ignition plug 16 stays within the flammable range for a relatively long time,
The ignition stability becomes extremely high. In addition, this improves the degree of freedom of the ignition timing control (for example, 40 ° CA before the compression top dead center (BTDC) of the cylinder 2 during stratified charge combustion operation)
(It can be set up to the range up to (TDC)), which can further improve fuel efficiency. Note that the graphs in FIG. 13 are all obtained by performing the test with the spread angle θ of the fuel spray set to θ = approximately 45 °.

Here, the inclination angle δ of the fuel spray center line F as described above will be examined in more detail. FIG. 15 shows that the setting angle of the injector 18, that is, the inclination angle δ of the spray center line F is either δ = 30 ° or δ = 36 °, and the spark plug 16 before the compression top dead center (TDC) of the cylinder 2 is set. The change in the local air-fuel ratio in the vicinity of the electrode was investigated. According to the figure, when δ = 36 °, more fuel adheres to the crown surface of the piston 5 than when δ = 30 °, so that the fuel is discharged relatively early to suppress this. I have to jet.

That is, as shown by the broken line in the figure, if the fuel injection is performed at 70 ° CA before the compression top dead center (BTDC), the local air-fuel ratio becomes sufficiently rich. To
If fuel injection is performed at BTDC 55 ° CA, the local air-fuel ratio becomes lean, and the period during which ignition stability can be ensured is slightly shortened. On the other hand, if δ = 30 °, the air-fuel ratio can be sufficiently enriched even when fuel injection is performed at BTDC 55 ° CA, as indicated by the solid line in FIG. = 30 ° and δ = 36
In comparison with the case of °, it can be said that it is preferable to set δ = 30 °.

Further, how the local air-fuel ratio at the ignition timing (for example, BTDC 30 ° CA) of the cylinder 2 when the engine 1 is in a predetermined operating state changes according to the change of the fuel injection timing. FIG. 16 shows the results of an investigation including the relationship with the inclination angle δ of the fuel spray center line F.
According to the figure, when δ = 30 °, even if the fuel injection timing is changed over a considerably wide range, the local air-fuel ratio near the electrode of the ignition plug 16 is generally within the flammable range.
It can be seen that the ignition stability is high even when the fuel injection timing is changed. On the other hand, when δ = 40 °, the air-fuel ratio becomes leaner than the flammable range in most cases.

Further, when δ = 20 °, the air-fuel ratio appears to be in an excessively rich state irrespective of the fuel injection timing. As shown by the dashed line in FIG. 13, the air-fuel ratio greatly changes with the progress of the crank angle, and the period during which the air-fuel ratio becomes excessively rich is extremely short. Note that FIG.
5. The test results shown in FIG. 16 are also obtained by setting the spread angle θ of the fuel spray by the injector 18 to θ = approximately 45 °.

(Setting of Spray Spread Angle) Subsequently, the spread angle θ of the fuel spray by the injector 18 is shown in FIG.
This will be described with reference to FIGS. The fact that the spray spread angle θ becomes large means that the fuel spray naturally diffuses and the air-fuel mixture is easily diluted. On the other hand, if the spray spread angle θ is too small, the vaporization and atomization of the fuel droplets and the mixing with the air are not sufficiently performed, and the mixture near the electrode of the ignition plug 16 may be in an excessively rich state. There is. Specifically, FIGS. 18 to 20 show the results of observing the distribution of the air-fuel mixture near the electrode of the ignition plug 16 with the spray spread angle θ being about 20 ° or about 60 °.
(a) is a view from the ceiling side of the combustion chamber 6 along the cylinder center line z, and (b) is a view from the injector 18 side in a direction orthogonal to the cylinder center line z. It is.

FIGS. 18 and 19 show the results of a test performed using a piston (see FIG. 39) in which a spherical concave portion is formed on the entire crown surface, which is different from this embodiment. According to the results, the spray spread angle θ = about 20 °
At the time of ignition of the cylinder, the air-fuel mixture is appropriately stratified in the vicinity of the electrode of the spark plug (indicated by a + sign in the figure), but when θ = approximately 60 °, the air-fuel mixture greatly expands, The ignition stability may be impaired due to a decrease in the local air-fuel ratio near the plug electrode.

On the other hand, according to the test results using a piston having a lemon-shaped recess provided on the crown surface as in this embodiment,
As shown in FIG. 20, the spray spread angle θ is set to θ = approximately 60 °.
The diffusion of the air-fuel mixture is suppressed, and θ =
It can be seen that the mixture is appropriately stratified as in the case of approximately 20 °. This is because the direction of the tumble flow is directed toward the fuel spray by the lemon-shaped concave portion of the piston crown surface, and the collision between the fuel spray and the tumble flow suppresses the diffusion of the fuel spray in the traveling direction. It is considered that the diffusion of the fuel spray in the width direction of the recess 5a is suppressed by the side wall surfaces on both sides in the width direction of the recess 5a and the squish flow flowing from the outside toward the inside of the recess 5a.

That is, as shown in FIG. 17, in this embodiment, the spread angle θ of the fuel spray from the injector 18 is set in the range of θ = approximately 20 ° to approximately 60 °, so that the electrode of the spark plug 16 The local air-fuel ratio in the vicinity can be stably set to a value within the flammable range. As described above, the spray divergence angle θ varies depending on the pressure state of the combustion chamber 6, but in the engine 1 of this embodiment, the spray divergence angle θ during the stratified combustion operation is set to θ =
More preferably, the value is in the range of approximately 40 ° to approximately 45 °.

(Position of Electrode of Spark Plug) As described above, in this embodiment, the behavior of the fuel spray by the injector 18 is controlled by the tumble flow T, and the flammable mixture having an appropriate concentration is retained in the center of the combustion chamber 6. That can be done. In order to reliably ignite the air-fuel mixture thus staying, as shown in FIGS. 9 and 21, the ignition plug 16 has its electrode extending along the cylinder center line z along the ceiling portion of the combustion chamber 6. It is arranged to protrude from.
That is, the electrode of the spark plug 16 is located above the fuel spray center line F from the injector 18 when viewed from the direction orthogonal to the cylinder center line z as shown in both figures (the one closer to the ceiling of the combustion chamber 6). ) Is located. In other words, when viewed along the spray center line F, the electrodes of the spark plug 16
With respect to the spray center line F, the combustion chamber 6
It is located on the side near the ceiling.

In this embodiment, the electrode of the spark plug 16 causes the injector 18 to inject fuel toward the diagonally lower part of the piston 5 crown surface. In this embodiment, the electrode is located below the injection hole of the injector 18 (the cylinder center line z). (In the direction of extension, closer to the crown surface of the piston 5). And, as described above, since the electrode of the ignition plug 16 is separated from the ceiling of the combustion chamber 6, the flame propagation is improved and the combustion state is improved.

Specifically, at the ignition timing of the cylinder 2 shown in FIG. 21 (for example, when the engine 1 is in a predetermined low load operation state and the ignition timing is BTDC 30 ° CA), the ceiling of the combustion chamber 6 When the distance from the cylinder center line z to the crown surface of the piston 5 is d, that is, the distance from the substantially uppermost position of the ceiling of the combustion chamber 6 to the deepest portion of the recess 5a of the piston 5 is d. The distance e (projection amount) on the cylinder center line z from the uppermost position of the ceiling of the combustion chamber 6 to the electrode of the ignition plug 16 is set to a value in the range of approximately 1 / 3d to approximately 2 / 3d. I have. Thus, in the compression stroke of the cylinder 2, the tumble flow flowing along the crown surface of the piston 5 is defined as a tumble positive flow Tm, while the tumble flow flowing along the ceiling of the combustion chamber 6 is defined as a tumble forward flow Ts. Further, after the start of fuel injection by the injector 18 and before the ignition timing of the cylinder at which the flow velocity of the positive tumble flow Tm is greatly attenuated by collision with the injected fuel, the electrode of the spark plug 16 It is located between Tm and the tumble forward flow Ts.

At the subsequent ignition point, the electrode of the ignition plug 16 is hardly affected by any of the tumble positive flow Tm and the tumble forward flow Ts, because the tumble positive flow Tm is significantly attenuated. become. That is, the electrode of the spark plug 16 is located away from the region where the main flow of the tumble forward flow Ts is generated, and is located on the center side of the vortex of the tumble flow T during the period after the fuel injection timing of the cylinder 2 and before the ignition timing. At the time, the mixture is maintained in a state in which the mixture easily stays.

The electrode of the ignition plug 16 is located above (including outside the spray area) the spray center line F of the geometrical spray area of the fuel spray from the injector 18 (see FIG. 9). Here, the geometrical spray area is an area of droplets of fuel spray assuming that there is no influence of tumble or swirl in the combustion chamber 6. By arranging the electrodes of the ignition plug 16 above the spray center line F in the geometrical spray area of the fuel spray, coarse fuel droplets contained in the initial spray from the injector 18 can be removed. It is possible to prevent the ignition plug 16 from smoldering, by avoiding sticking to the electrodes.

FIG. 22 shows that, while changing the protruding amount e of the electrode of the ignition plug 16, the BT of the cylinder 2 is positioned at the position of the electrode.
This is a measurement of the flow velocity of the tumble flow T immediately before fuel injection near DC 55 ° CA. According to the figure, the flow velocity of the tumble flow T becomes minimum near the protrusion amount e = 1１d, and even if the protrusion amount e is smaller or larger than that, the tumble flow changes with the change of the protrusion amount e. The flow velocity of T gradually increases. This means that near 55 ° CA BTDC,
This means that the electrode of the ignition plug 16 is located near the center of the vortex of the tumble flow T in the vicinity of the protrusion amount e = １d.

FIG. 23 shows such a spark plug 1
6 shows a change in a fluctuation rate (a value serving as an index indicating combustion fluctuation) of the indicated mean effective pressure (PI) of the engine 1 when the protrusion amount e of the electrode 6 is changed. According to FIG.
This Pi fluctuation rate exceeds a predetermined reference line (for example, about 5%) when the protrusion amount e <0.2d. In this state, the local air-fuel ratio near the electrode of the ignition plug 16 is lean, It can be seen that the flammability has decreased. On the other hand, if the protrusion amount e ≧ 0.2d, the larger the protrusion amount e is, the closer the electrode of the ignition plug 16 is to the crown surface of the piston 5 and the local air-fuel ratio near the electrode becomes rich. Thus, the combustion fluctuation rate becomes smaller. However, when the protrusion amount e is increased beyond 0.5 d, the electrode of the ignition plug 16 comes into contact with the spray center line F in the geometrical spray area of the fuel spray, and the initial spray There is a possibility that a problem may occur due to the attachment of coarse fuel droplets.

In short, the protruding amount e of the electrode of the ignition plug 16 is optimal in view of the easiness of stagnation of the air-fuel mixture, the ignitability of the air-fuel mixture, and the prevention of the attachment of coarse fuel droplets. In the vicinity of the BTDC 55 ° CA shown in FIG. 23, the range e is approximately 0.2 d as indicated by an arrow in FIG.
~ 0.5d.

Here, in general, the distance d on the cylinder center line z from the ceiling of the combustion chamber 6 to the crown of the piston 5 in the compression stroke of the cylinder 2 becomes geometrically smaller as the piston 5 rises. Assuming that the amount of protrusion e of the electrodes of the spark plug 16 is constant, their ratio e / d changes as shown in FIG. Therefore, in this embodiment, the value of the protrusion amount e is set such that the value of the ratio e / d of the protrusion amount is a value of the region sandwiched between the two graphs shown in FIG. According to the figure, for example, BT which is the most advanced injection start timing (crank angle position) in the stratified combustion region
At DC 140 ° CA, e / d is approximately 10% to approximately 20%.

The BTDC 55 ° which is the injection start timing (crank angle position) in the relatively low rotation speed region of the stratified combustion region.
In CA, e / d is approximately 20% to approximately 50%. further,
It is the injection start timing (crank angle position) in the substantially idle range.
At BTDC40 ° CA, e / d = approximately 25% to approximately 60%, and at the compression top dead center (TDC) of the cylinder 2, the protruding ratio e / d of the spark plug electrode is e / d = approximately 40% to about 9
It is 5%. In this embodiment, the ignition timing of each cylinder 2 in the stratified combustion region ranges from approximately BTDC 40 ° CA to approximately TDDC.
C is set in the range.

(Adjustment of Tumble Velocity and Spray Penetration Force) Finally, in this embodiment, the operating state of the engine 1 is changed by adjusting the penetration force of the fuel spray by the injector 18 according to the tumble flow velocity. Therefore, a stable mixture can be formed. Hereinafter, this point will be described in detail.

First, FIG. 25 shows that the flow field of the combustion chamber 6 near the ignition timing of the cylinder 2 is adjusted by the CFD (computational fluid dynamics) by adjusting the penetration force of the fuel spray by the injector 18 so as to correspond to the velocity of the tumble flow. ) Was analyzed by applying As shown by the thick arrows in the figure, the tumble flow and the spray flow flow from the left and right sides of the figure toward the center side along the piston crown surface, respectively, and their collision points A are indicated by + signs in the figure. In the vicinity of the electrode of the spark plug. Thereby, for example, as shown in FIG.
In the vicinity of the ignition timing of the cylinder 2, a combustible mixed air mass having an appropriate concentration state is supplied to an electrode of the ignition plug 16 (indicated by a + mark in the figure)
It becomes possible to stay near.

Here, in general, the flow rate of the tumble flow in the combustion chamber changes according to the rotation speed of the engine (the rotation speed of the crankshaft 7), so that the mixture is appropriately stratified over a wide operating range. In this case, it is necessary to adjust the fuel spray penetration force by the injector according to the operating state of the engine. For this purpose, for example, it is conceivable to change and adjust the fuel injection pressure by the injector. At this time, for example, when the engine is idling, the flow velocity of the tumble flow becomes too low.Therefore, if the injection pressure of the fuel is reduced to correspond to this, the atomization characteristics of the fuel may be reduced. is there. Conversely, for example, when the flow velocity of the tumble flow exceeds a predetermined upper limit in the middle and high rotation range of the engine, if the fuel injection pressure is increased to correspond to this, the collision between the fuel spray and the tumble flow becomes severe. There is a problem that too much. In this case, although the air-fuel mixture stays in the vicinity of the electrode of the ignition plug 16, for example, as shown in FIG. It may be damaged.

On the other hand, in this embodiment, the opening degree of the intake air flow control valve 34 (see FIG. 3) disposed at the intake port 10 of each cylinder 2 is adjusted mainly in accordance with the change in the rotation speed of the engine 1. Then, the flow rate of the tumble flow T in the combustion chamber 6 is changed, and the injector 18
By appropriately adjusting the fuel injection pressure, the stratified combustion region (a) (see FIG. 6) achieves appropriate stratification of the air-fuel mixture.

More specifically, as shown in FIG. 27 (a), when the engine 1 is in the stratified charge combustion region and its rotation speed is equal to or lower than the first set value ne1 (for example, 2500 rpm), the intake flow control valve 34 is fully turned off. Closed. By doing this,
The flow velocity of the tumble flow is increased as compared with the state in which the intake flow control valve 34 is open (indicated by a phantom line in the figure). Next, if the engine rotational speed exceeds the first set value ne1 and the flow velocity of the tumble flow reaches a predetermined upper limit value, the flow velocity of the tumble flow T does not increase further even if the engine rotational speed increases. Then, the intake flow control valve 34 is gradually opened according to the increase in the engine speed. And
When the engine speed is equal to the second set value ne2 (for example, 3500r
pm), the engine 1 shifts to the uniform combustion region, thereafter the intake flow control valve 34 is fully opened,
The intake air volume is secured.

The operation of the intake flow control valve 34 is controlled by the ECU 50 in accordance with a predetermined control program.
Corresponds to a tumble flow rate control means that operates the intake flow control valve 34 so as to suppress an increase in the tumble flow rate corresponding to the increase in the rotation speed when the rotation speed of the engine becomes equal to or higher than the set value.

Then, the fuel injection pressure by the injector 18 is changed so as to correspond to the above-described change in the tumble flow velocity, and the penetration force of the fuel spray is changed as shown in FIG. I have. That is, when the engine 1 is in the stratified combustion region and the rotation speed thereof
At ne1 or less, the penetration force of the fuel spray is increased in accordance with the increase in the rotation speed. On the other hand, from when the engine speed exceeds the first set value ne1 until the engine speed reaches the second set value ne2,
The penetration force of the fuel spray is maintained substantially constant. Further, if the rotation speed exceeds the second set value ne2 and the engine 1 shifts to the uniform combustion region, thereafter, the fuel injection pressure is determined according to the balance between the fuel injection amount and the time interval during which the fuel can be injected. Like that.

The adjustment of the spray penetration force is performed by the ECU 5
0 according to a predetermined control program.
Is activated, the fuel pressure (fuel pressure) supplied to the injector 18 is changed. FIG. 28 shows the relationship between the fuel pressure and the spray penetration force.
As shown in FIG. 27 (b), the relationship between the engine rotation speed and the spray penetration force is as shown in FIG. Therefore, the fuel injection pressure is adjusted. Accordingly, the ECU 50 increases the penetration force of the fuel spray by the injector 18 in accordance with the increase in the engine speed until the engine speed reaches the set value, while increasing the penetration force when the engine speed becomes equal to or higher than the set value. This also corresponds to the configuration of the fuel injection control means for suppressing the increase in the fuel injection.

The tumble flow velocity and the spray penetration force may be adjusted as shown in FIG. 29, for example. That is,
As shown in (a), the rotation speed of the engine 1 is reduced to a second set value n.
During e2 or less, the intake flow control valve 34 is gradually opened in accordance with the increase in the engine rotation speed, so that the flow velocity of the tumble flow T is gradually increased within the range of the upper limit value or less. Then, the penetration force of the fuel spray may be gradually increased in accordance with the engine speed.

As described above, in the direct injection engine 1 of this embodiment, the mixture convective near the electrode of the ignition plug 16 is prevented from being greatly diffused and diluted.
The opening degree of the intake flow control valve 34 is controlled so that the flow rate of the tumble flow T in the combustion chamber 6 does not exceed the upper limit. This is because the combustion in the stratified combustion region of the engine 1 is performed. This means that the tumble ratio of the chamber 6 falls within a predetermined range. That is, in this embodiment, the upper limit value of the tumble flow velocity is set so that the tumble ratio of the combustion chamber 6 becomes a value in a range of about 1.1 to about 2.3,
It can be said that the rotational speeds ne1 and ne2 of the engine 1 are set to correspond to this. Here, the tumble ratio is a measure of the strength of the tumble flow in the in-cylinder combustion chamber.Specifically, the vertical angular velocity of the intake flow in the cylinder is measured for each valve lift of the intake valve. , And then defined as a value obtained by dividing the integrated value by the engine angular velocity. Therefore, if the tumble ratio is constant, the flow rate of the tumble flow in the cylinder increases as the engine rotation speed increases.

The measurement of the intake air flow longitudinal angular velocity is performed by, for example, an apparatus 80 having a structure as shown in FIG. That is, in the figure, reference numeral 4 denotes a cylinder head of the engine, which is installed laterally in the rig device 80, and is measured by an intake supply passage 81 at an upstream end of the intake ports 10, 10 of a predetermined cylinder. The intake air supply device 82 is connected. On the other hand, the downstream ends of the intake ports 10 and 10 are connected to a substantially central portion of a measurement tube 84 by a connection pipe 83. The intake supply device 82 adjusts the air supplied from the blower 85 so that the pressure difference between the intake ports 10 and 10 and the measurement tube 84 is close to the atmospheric pressure when the throttle is fully opened. Port 10, 10
Is to be sent to

Further, the measuring tube 84 is formed of a cylinder having the same diameter as, for example, the diameter of a cylinder and having a length of about 10 times that of the cylinder, and a honeycomb-shaped rotor 86 at the upper end in the figure.
Is connected, and a dummy rotor 88 having the same rotational resistance as the rotor 86 is connected to the lower end of the figure. The length of the measuring tube 84 is set to about 10 times its diameter to ensure the accuracy and stability of the measurement. For the same reason, the connecting tube 83 has a length h. It is short (eg, about 2 cm) and has the same diameter as the cylinder.

When the air from the blower 85 is supplied to the intake ports 10 and 10 through the intake supply passage 81, the intake air flows through the intake ports 10 and 10 into the inside of the measurement tube 83. The flow then flows into a swirling flow that rotates in the circumferential direction and moves to both ends of the measurement tube 84, and a rotational force is applied to the rotor 86 by the swirling flow. This rotational force is measured via a torque arm 89 of the impulse meter 87, and the intake air flow longitudinal angular velocity is determined based on the measured value.

In general, the penetration force of the fuel spray from the injector 18 changes depending on the temperature of the combustion chamber 6 of the engine 1. The temperature of the combustion chamber 6 depends on the temperature of the engine 1 at that time. And changes depending on the warm-up state of the engine 1, the presence or absence of exhaust gas recirculation (EGR), and the like. For example, after the engine is warmed up, the temperature of the intake air drawn into the combustion chamber 6 is higher than before the warm-up is completed, the combustion temperature is also increased, and the temperature of the combustion chamber 6 is higher. Further, since the exhaust gas temperature is also increased by the increase of the combustion temperature, when the exhaust gas is recirculated by the EGR passage 43, the intake air temperature is increased by the influence of the high-temperature exhaust gas. As a result, for example, as shown by the solid line in FIG. 31 (a), when the temperature state of the combustion chamber 6 increases, the vaporization of the fuel is promoted, and the spray penetration force tends to decrease as shown by the broken line in FIG. is there.

On the other hand, it is extremely important to balance the strength of the tumble flow and the penetration force of the fuel spray for the appropriate stratification of the air-fuel mixture as described above. After setting the fuel injection pressure by the fuel injection amount 18 and the engine rotation speed, the fuel injection pressure is further corrected according to the temperature state of the combustion chamber 6 as described above. Specifically, as shown by the solid line in FIG. 31 (b), the fuel injection pressure (fuel pressure)
The fuel injection pressure is corrected to be higher as the temperature state of the engine 1 is estimated to be higher, based on the load state, rotation speed, engine water temperature, presence or absence of exhaust gas recirculation by the EGR valve 44, etc. As a result, the spray penetration force increases as shown by the broken line in FIG. That is, the combustion chamber 6 in the cylinder 2
Even if the temperature of the fuel cell rises and the spray penetration force decreases, the injection pressure is increased and the spray penetration force is kept constant, whereby the concentration state of the air-fuel mixture near the electrode of the ignition plug 16 fluctuates. Can be suppressed.

(Operation and Effect) Next, the operation and effect of the spark ignition type direct injection engine 1 of the present invention during stratified combustion will be described.

When the engine 1 is in the stratified combustion region (a), when the piston 5 descends from the top dead center position in the intake stroke of the cylinder 2 shown in FIG. 32, the head portion of the intake valve 12 and the intake port are opened. The intake air flows into the combustion chamber 6 from the gap with the opening end of the cylinder 10, and a tumble flow T is generated as shown by an arrow in FIG. Specifically, the intake air sucked into the combustion chamber 6 by the lowering of the piston 5 is mainly supplied to the intake port 10.
Flows into the combustion chamber 6 from a portion near the spark plug 16 at the open end of the combustion chamber 6. Then, as the piston 5 further descends, it moves downward along the cylinder inner peripheral surface on the exhaust side (left side in the figure), and is then bent to the intake side (right side in the figure) along the crown surface of the piston 5. Thus, the tumble flow T flows further upward from there and turns largely vertically in the entire combustion chamber 6.

Subsequently, when the cylinder 2 moves to the compression stroke and the piston 5 rises from the bottom dead center position, the volume of the combustion chamber 6 decreases with the rise of the piston 5, so that the tumble flow T is crushed. Although it becomes compact and its flow velocity gradually decreases, the tumble flow T does not collapse,
The cylinder 2 is held until the middle stage of the compression stroke. Further, even after the middle stage of the compression stroke of the cylinder 2,
Since a combustion chamber space of an appropriate shape is left between the pent roof type combustion chamber 6 ceiling and the recess 5a of the piston 5 crown surface, the preservability of the tumble flow T in the combustion chamber 6 is improved. . At this time, the tumble flow T flowing from the exhaust side to the intake side (from left to right in the figure) along the crown surface of the piston 5
The (tumble normal flow) turns around near the injection hole of the injector 18, and then flows from the intake side to the exhaust side along the ceiling of the combustion chamber 6.
Is guided by the concave portion 5 a of the crown surface of the piston 5 and flows toward the injection hole of the injector 18.

For this reason, as shown in FIG. 11, when fuel is injected by the injector 18, most of the fuel spray has a strong flow of the tumble flow T flowing along the concave portion 5a of the piston 5 crown surface. However, the vehicle will collide so as to substantially face. As a result, vaporization and atomization of the fuel droplets and mixing with the surrounding air are promoted, and the fuel spray is gradually decelerated while proceeding so as to displace the tumble flow T, so that the cylinder 2 shown in FIG. At the ignition timing, a combustible mixture having an appropriate concentration is formed, and this combustible mixture stays around the electrode of the ignition plug 16 at the center of the combustion chamber 6. Moreover, the squish flow flowing from the outside of the recess 5a toward the center of the cylinder suppresses the diffusion of the combustible air-fuel mixture, thereby increasing the stagnation accuracy of the spark plug 16 around the electrode. Then, when the ignition plug 16 is energized, the mixture is ignited, and favorable stratified combustion is performed. Then, when the ignition plug 16 is energized, the mixture is ignited, and favorable stratified combustion is performed.

That is, in the compression stroke of the cylinder 2, the tumble flow T is guided by the concave portion 5 a of the crown surface of the piston 5 and flows from the center of the combustion chamber 6 toward the injection hole of the injector 18. In contrast, by effectively colliding the fuel spray from an appropriate direction and in an appropriate spread state, the air-fuel mixture can be appropriately stratified and retained in the center of the combustion chamber 6.

At this time, the opening degree of the intake flow control valve 34 provided in the intake port 10 is controlled in accordance with the operating state of the engine 1 so that the flow rate of the tumble flow T is adjusted to be within an optimum range. Similarly, the fuel injection pressure from the injector 18 is controlled, and the penetration force of the fuel spray is adjusted so as to fall within an optimum range corresponding to the flow velocity of the tumble flow T. As a result, even if the operating state of the engine 1 changes from the idling region to the medium load rotation region in the stratified combustion region (a), regardless of the change in the operating state, the strength of the tumble flow and the penetration of the fuel spray are maintained. Are balanced in the optimal range not too strong and not too weak, respectively.
As described above, the combustible air-fuel mixture having the optimum concentration can be retained in the center of the combustion chamber 6.

Therefore, according to the spark ignition type direct injection engine 1 according to this embodiment, the behavior of the fuel spray is appropriately controlled by the tumble flow in the combustion chamber 6 of the cylinder 2 over the entire stratified combustion region (a). As a result, optimal stratification of the air-fuel mixture can be achieved irrespective of a change in the operating state, and fuel economy and output can be improved by realizing good stratified combustion. Moreover, by such a method of forming an air-fuel mixture,
Since a favorable stratified combustion state can be realized up to a higher load or a higher rotation range than in the past, a higher fuel efficiency improvement effect can be obtained in the entire operation region of the engine 1 by expanding the stratified combustion region. .

Further, as described above, the mixture in an appropriate concentration state can be retained near the electrode of the ignition plug 16 at the ignition timing of the cylinder 2, so that the mixture can be stably ignited. The fuel consumption and the output can be improved even when the period is extremely long and the degree of freedom in controlling the ignition timing of the cylinder 2 is increased.

A conventional example (Japanese Unexamined Patent Application Publication No. 11-161338)
), The inner wall surface of the cavity facing the injector 18 is located near the electrode of the ignition plug, and does not hinder the growth of the flame nucleus in the early stage of combustion or reduce the propagation of the flame. In addition, since the adhesion of fuel to the piston crown surface can be reduced, fuel efficiency and output can be further increased, and the concentration of unburned HC in exhaust gas can be reduced.

More specifically, FIGS. 33 and 34 show the effects of improving fuel efficiency and purifying exhaust gas of the spark ignition type direct injection engine 1 of this embodiment in comparison with the conventional direct injection engine. It is. That is, FIG. 33 shows the fuel efficiency improvement rate, fuel efficiency rate, and HC emission rate of the engine 1 at the time of low-speed operation (for example, 1500 rpm), respectively.
This is based on a so-called port injection type engine in which fuel is injected into an intake port. As shown by the solid line in the figure, in the direct injection engine 1 of the present invention,
It can be seen that even in the low rotation speed region, the fuel efficiency is improved as compared with the conventional direct injection engine indicated by the broken line in the same figure, and the emission of unburned HC can be significantly reduced. This is considered to be because the amount of fuel adhering to the crown surface of the piston can be reduced.

FIG. 34 shows a similar test result at the time of medium rotation operation of the engine 1 (for example, 2500 rpm). In the present invention shown by a solid line in FIG. It can be seen that the effect of reducing the emission of fuel HC is great. This is because the direct injection engine 1 of the present invention appropriately stratifies the air-fuel mixture around the electrodes of the ignition plug even in an operation region where the fuel spray cannot be captured in the piston cavity in the conventional direct injection engine. This is because good stratified combustion can be realized.

(Embodiment 2) FIGS. 35 to 39 show the shape of a piston 5 of a spark ignition type direct injection engine 1 according to Embodiment 2 of the present invention. Are provided with concave portions 5a of different shapes instead of the lemon-shaped concave portions 5a as in the first embodiment. The configuration of the engine 1 according to the second embodiment is the same as that according to the first embodiment except for the shape of the concave portion 5a. Therefore, the same members will be denoted by the same reference numerals, and description thereof will be omitted.

In the piston 5 shown in FIGS. 35 (a) and 35 (b), the portion of the recess 5a on the intake side (right side in the figure) has substantially the same shape as the lemon-shaped recess of the first embodiment. ,
In the portion on the exhaust side (left side in the figure), the width of the concave portion 5a is substantially constant from the substantially central portion in the longitudinal direction of the concave portion 5a to the end portion on the exhaust side. That is, in this case, the concave portion 5a
Of the ends on both sides in the length direction, the width of the end farther from the injector 18 (exhaust side) is substantially the same as the width of a substantially central portion in the length direction of the concave portion 5a. On the intake side of the crown of the piston 5, a relatively small and shallow recess 5b is formed so as to be continuous with the recess 5a in order to avoid interference with fuel spray from the injector 18.

Also, in the piston 5 shown in FIGS. 36 (a) and (b), similarly to the piston 5 shown in FIG. 35, the portion of the recess 5a on the intake side (right side in the figure) is substantially the same as that of the embodiment. 1 has the same shape as the lemon-shaped concave portion, but the portion on the exhaust side (the left side in the figure) has a width of the concave portion 5a from a substantially central portion in the longitudinal direction of the concave portion 5a toward the end on the exhaust side. Is gradually expanding. In other words, in this case, the end farthest from the injector 18 (exhaust side) among the ends on both sides in the longitudinal direction of the concave portion 5a.
Is wider than the width of a substantially central portion in the longitudinal direction of the concave portion 5a.

In the piston 5 shown in FIG. 35 or 36, during the compression stroke of the cylinder 2, the combustion chamber 6
The tumble flow T flowing from the ceiling to the crown surface of the piston 5 along the inner peripheral surface of the cylinder 2 is smoothly introduced into the recess 5a without much attenuation. The flow velocity of the tumble flow T for controlling the behavior of the fuel spray from the injector 18 can be sufficiently ensured.

More specifically, FIG. 37 shows the relationship between the change in the crank angle and the change in the tumble ratio during the compression stroke of the cylinder 2. According to FIG. As compared with the case where a lemon-shaped concave portion as shown in FIG. 1 is provided (indicated by a broken line in the drawing), a concave portion having a substantially constant width on the exhaust side is provided as shown in FIG. Also, in the case of providing a concave portion whose width on the exhaust side gradually increases as shown in FIG. 36 (shown by a solid line in the figure), a large tumble ratio is always obtained during the compression stroke. That is, if the piston 5a as in the second embodiment is used, the flow rate of the tumble flow T can be easily ensured even when the rotation speed of the engine 1 is particularly low. Thus, the opening degree of the intake flow control valve 34 can be controlled to be relatively large, whereby the loss due to the intake throttle of the engine 1 is reduced, and the fuel consumption is further reduced.

Further, the recess 5a of the piston 5 shown in FIG. 38 has a lemon shape similar to that of the first embodiment when viewed along the cylinder center line z as shown in FIG. As shown in FIG. 2B, when viewed perpendicularly to the cylinder center line z and along the axis of the crankshaft 7 of the engine 1, a portion closer to the exhaust side than the center of the cylinder is closer to a portion closer to the intake side. Is also formed to be deeper. That is, in this case, the volume of the concave portion 5a is smaller on the intake side closer to the injector 18 than the exhaust side on the opposite side to the cylinder center line z, so that the temperature state is relatively low. Further, the volume of the combustion chamber on the intake side, where the propagation of the flame tends to be slow, is made smaller than that on the exhaust side, so that the occurrence of abnormal combustion on the intake side can be prevented and knocking can be suppressed.

In addition, the piston 5 shown in FIG.
Spherical recess 5a extending over the entire crown surface of piston 5
Is provided. In order to avoid interference with the fuel spray from the injector 18 on the intake side of the crown surface of the piston 5 as shown in FIGS.
A relatively small and shallow concave portion 5b is formed, and the combined volume of the two concave portions 5a and 5b is determined based on the compression ratio of the cylinder 2. And this piston 5
In the case of the above, the combustion chamber 6 depends on the shape of the recess 5a as described above.
(Surface Volume Ratio) and the uniformity of flame propagation in all directions, up, down, left, and right when viewed along the cylinder center line z. Output is improved.

The recess 5 of the piston 5 shown in FIG.
In a, there is a tendency that the fuel spray from the injector 18 is more likely to spread to both left and right sides as viewed along the spray center line F, as compared with other examples. That is, in the first embodiment, the tumble flow T is guided inward by the side walls on both sides in the width direction of the lemon-shaped concave portion 5a, and the squish flows from the side walls on both sides and the outside thereof cause the fuel spray to be sprayed. Although the diffusion is suppressed, it is difficult to obtain the effect of the side wall surface of the concave portion and the squish flow in the one shown in FIG. 39, and therefore, in order to appropriately stratify the air-fuel mixture, the first embodiment is used. In comparison, the spread angle θ of the fuel spray by the injector 18 is relatively small (for example, θ = about 20 ° to about 5).
0 °).

(Embodiment 3) FIG. 40 shows the shape of the fuel spray from the injector 60 in the spark ignition type direct injection engine 1 according to the third embodiment of the present invention. In a virtual plane orthogonal to the center line F, the size of the fuel spray becomes flat in one of the vertical and horizontal directions. The configuration of the engine 1 of the third embodiment is the same as that of the first and second embodiments except for the configuration of the injector 60, and therefore, the same members will be denoted by the same reference numerals and description thereof will be omitted. .

The injector 60 according to the third embodiment is a swirl injector in which a velocity component in the swirling direction is added to the fuel injected from the injection hole at the tip end similarly to the first embodiment. Although not shown, the injection hole is shifted to one side with respect to the axis J of the injector 60, and the shape of the injection hole is a flat slit shape. As a result, when the injector 60 is viewed from above as shown in FIG. 3A, the shape of the fuel spray from the injector 60 is greatly expanded as compared with the injector 18 of the first embodiment, and the relatively large spray In this case, the injector 6 has a divergence angle θ1, as shown in FIG.
When viewed from the side of 0, the fuel spray has a relatively small spread angle θ2.

In the third embodiment, the flat spray injector 60 described above has a fuel spray spread angle along the axis of the crankshaft 7 when the fuel spray spread angle is viewed along the cylinder center line z. It is installed so that it is larger than it looks. By doing so, when viewed along the axis of the crankshaft 7 as shown in FIG. 41, the spread angle θ2 of the fuel spray in the vertical direction of the drawing where the cylinder center line z extends becomes relatively small. As a result, more fuel can effectively collide with the positive tumble flow Tm flowing along the concave portion 5a of the piston 5. In addition, since the amount of fuel spray that spreads toward the ceiling of the combustion chamber 6 is small, the influence of the tumble forward flow Ts on the fuel spray is reduced, and the amount of fuel spray that spreads toward the crown of the piston 5 is also small. Fuel adhesion to the surface can also be suppressed.

Therefore, according to the third embodiment, by using the injector 60 in which the fuel spray becomes flat, a larger amount of fuel is colliding with the tumble positive flow Tm more effectively than in the first embodiment. This makes it possible to improve the accuracy of fuel spray behavior control by the tumble positive flow Tm. Moreover, it is easy to position the electrode of the ignition plug 16 outside the geometric fuel spray area by using the above spray shape,
In this way, the adhesion of the fuel droplets to the electrode can be suppressed as much as possible, and the occurrence of smoldering of the ignition plug 16 can be reliably prevented.

(Embodiment 4) FIG. 42 shows the shape of fuel spray from an injector 61 in a spark-ignition direct injection engine 1 according to Embodiment 4 of the present invention. The asymmetric spray is larger on one side of the spray center line F than on the other side. The configuration of the engine 1 of the fourth embodiment is the same as that of the first to third embodiments except for the configuration of the injector 61, and therefore, the same members will be denoted by the same reference numerals and description thereof will be omitted. .

A feature of the injector 61 of the fourth embodiment lies in the structure of the injection hole portion shown enlarged in FIG. That is, in the injector 61, since the nozzle 62 of the injection hole portion is formed to be inclined with respect to the axis J of the injector 61, the nozzle 62 is supplied from the upstream side through the seat portion between the nozzle 62 and the ball valve 63. The direction in which the fuel travels is bent in the direction in which the nozzle 62 deflects as shown by the arrow in FIG. 3B, and the fuel flow rate on the deflection side increases. As a result, as shown in FIG. 7A, assuming that the penetration force of the fuel spray on the deflection side of the nozzle 62 is L1 and the penetration force of the fuel spray on the opposite side is L2, the fuel on the deflection side is as described above. Since the flow rate ratio increases, a deviation occurs in the momentum of the fuel spray, and the penetration force L1 of the fuel spray on the deflection side becomes larger than L2.

In the fourth embodiment, as shown in FIG. 43, the injector 61 of asymmetrical spraying as described above uses the fuel spray penetrating force from the injector 61 during the compression stroke of the cylinder 2 to move the fuel spray center. The fuel injection direction of the injector 61 is determined such that the inclination angle δ of the spray center line F is set to be larger on the side closer to the crown surface of the piston 5 (the lower side in the figure) than on the opposite side with respect to the line F. This is set to be slightly smaller than in the first embodiment.

Therefore, according to the fourth embodiment, of the fuel spray from the injector 61, the portion of the fuel spray near the crown surface of the piston 5 collides substantially directly with the tumble positive flow Tm. Even in a state where the spread angle of the fuel spray is relatively large, a sufficiently large amount of fuel can effectively collide with the tumble positive flow Tm. On the other hand, since the penetration of the fuel spray on the side close to the ceiling of the combustion chamber 6 is small, even if the fuel spray is caused to flow by the tumble forward flow Ts, the adverse effect caused by the fuel spray is extremely small, and the over-diffused mixture is Because the amount can be reduced, stratification improves,
Fuel efficiency is improved. Therefore, according to the fourth embodiment, it is possible to improve the accuracy of the fuel spray behavior control by the tumble positive flow Tm as in the third embodiment.

Further, as shown in FIG. 44, the asymmetric spray injector 61 is connected to the fuel spray penetration force in the compression stroke of the cylinder 2 on the side closer to the ceiling of the combustion chamber 6 with respect to the spray center line F. May be arranged to be larger than the opposite side. In this case, the fuel injection direction of the injector 61 is set to the inclination angle δ of the spray center line F in order to cause the side of the fuel spray near the ceiling of the combustion chamber 6 to collide with the tumble positive flow Tm substantially. Is preferably set to be larger than those in the above embodiments. In this manner, similarly to the third embodiment, even in a state where the spread angle of the fuel spray from the injector 61 is relatively large, it is possible to cause a sufficiently large amount of fuel to effectively collide with the tumble positive flow Tm, On the other hand, since the fuel spray on the side close to the crown surface of the piston 5 has a small penetration force, the amount of fuel attached to the crown surface of the piston 5 does not increase so much. The accuracy of behavior control can be improved.

In this case, the penetration force of the fuel spray from the injector 18 on the side close to the ceiling of the combustion chamber 6 increases, so that the electrode of the ignition plug 16
Preferably, it is located outside the geometric fuel spray area in the compression stroke of the cylinder 2.

(Other Embodiments) The present invention relates to the first embodiment.
The present invention is not limited to the above-described embodiments, but includes other various embodiments. That is, in each of the above-described embodiments, the opening degree of the intake flow control valve 34 provided in the intake port 10 is changed to adjust the flow rate of the tumble flow T in the combustion chamber 6, but is not limited thereto. For example, the flow rate of the tumble flow T may be adjusted by changing the opening / closing operation timing of the intake valve 12 or the exhaust valve 13 by the variable valve mechanism 14. That is, if the closing operation timing (valve timing) of the exhaust valve 13 is retarded as shown by a solid line in FIG. The amount of burned gas flowing back into the chamber 6 increases, whereby the flow rate or flow rate of the intake air drawn into the combustion chamber 6 decreases, and the flow rate of the tumble flow T decreases. In addition, since the temperature of the combustion chamber 6 is increased by the backflow of the burned gas, even if the injection pressure of the fuel is reduced so as to correspond to the decrease in the tumble flow velocity, the vaporization and atomization characteristics of the fuel deteriorate. I will not do it.

As shown by the solid line in FIG. 11B, if the opening and closing operation timings of both the intake valve 12 and the exhaust valve 13 are both retarded, the intake efficiency is lowered in addition to the above-mentioned operation. As a result, the flow velocity of the tumble flow T decreases. By doing so, the amount of change in the opening / closing operation timing of the intake valve 12 and the exhaust valve 13 can be relatively small, and fluctuations in the operating state of the engine 1 due to this can be suppressed. Alternatively, although not shown, the flow rate of the tumble flow T can be reduced by retarding only the opening / closing operation timing of the intake valve 12.

[0138]

As described above, according to the spark ignition type direct injection engine according to the first aspect of the present invention, a concave portion is formed on the crown surface of the piston so as to follow the tumble flow, and the fuel is supplied to the tumble flow. By optimally setting the injection direction and the position of the electrode of the spark plug, during stratified charge combustion operation, the fuel spray injected by the fuel injector during the compression stroke of the cylinder is not confined in the cavity of the piston, etc. It can be appropriately stratified around. This makes it possible to appropriately stratify the air-fuel mixture around the spark plug regardless of the operating condition of the engine. When the fuel efficiency can be greatly improved. In addition, since the mixture can be retained near the electrode of the spark plug as described above while suppressing the fuel from adhering to the piston crown surface, the degree of freedom of the ignition timing control is increased, which also increases the fuel consumption and output. Is improved.

According to the second or third aspect of the present invention, the positions of the electrodes of the ignition plug are embodied, respectively.
The effect of the invention can be sufficiently obtained.

According to the fourth aspect of the present invention, by appropriately setting the direction of fuel injection by the fuel injection valve with respect to the electrode position of the spark plug, the fuel spray is reliably retained near the electrode. The effect of the invention of Item 1 can be sufficiently obtained.

According to the fifth aspect of the invention, by appropriately setting the direction of fuel injection by the fuel injection valve, the fuel spray can be made to effectively collide with the positive tumble flow while reducing the influence of the tumble forward flow. The behavior of the fuel spray can be more accurately controlled by the tumble positive flow.

According to the sixth aspect of the present invention, the shape of the fuel spray by the fuel injection valve is made flat, and the spread angle of the fuel spray in the direction in which the cylinder center line extends is viewed relative to the axis of the crankshaft. By reducing the effect of the tumble forward flow on the fuel spray by making the fuel spray smaller, a sufficient amount of fuel is made to effectively collide with the tumble forward flow,
The accuracy of the fuel spray behavior control by the tumble positive flow can be improved. Further, the adhesion of the fuel droplets to the electrode of the ignition plug can be suppressed as much as possible, and the occurrence of smoldering can be prevented.

According to the seventh aspect of the present invention, the penetration force of the fuel spray by the fuel injection valve is asymmetrical with respect to the spray center line so as to be relatively large on the side close to the crown surface of the piston, and the side on which the penetration force is large. By colliding the fuel spray with the tumble forward flow substantially in the opposite direction, it is possible to reduce the adverse effect of the tumble forward flow on the fuel spray while effectively colliding a sufficiently large amount of fuel with the tumble forward flow. The accuracy of the behavior control of the fuel spray by the tumble positive flow can be improved.

According to the eighth aspect of the present invention, the penetration force of the fuel spray by the fuel injection valve is asymmetric so as to be relatively large with respect to the spray center line on the side close to the combustion chamber ceiling, and the side having the greater penetration force. By causing the fuel spray to substantially collide with the tumble positive flow, it is possible to effectively cause a sufficient amount of fuel to collide with the tumble positive flow while suppressing the fuel spray from adhering to the piston crown surface. Thus, the accuracy of the behavior control of the fuel spray by the tumble positive flow can be improved. Further, the adhesion of the fuel droplets to the electrode of the ignition plug can be suppressed as much as possible, and the occurrence of smoldering can be prevented.

According to the ninth aspect of the present invention, by making the shape of the concave portion in the crown face of the piston appropriate, the tumble flow flowing along the concave portion can stably collide with the fuel spray. Thus, diffusion of the fuel spray in the width direction of the concave portion can be suppressed, so that appropriate stratification of the air-fuel mixture can be promoted.

According to the tenth or eleventh aspect of the present invention, the shape of the concave portion in the ninth aspect of the present invention is made more appropriate so that the tumble flow is smoothly introduced into the concave portion and the flow velocity thereof is sufficiently ensured. Can be.

According to the twelfth aspect of the invention, by forming a squish area at least on the exhaust side of the piston crown surface, the squish flow generated near the ignition timing of the cylinder reduces the diffusion of the air-fuel mixture at the center of the combustion chamber. By suppressing the mixture, the air-fuel mixture can stay longer near the electrode of the spark plug.

According to the thirteenth aspect of the present invention, the volume of the concave portion is made smaller on the intake side where the temperature is relatively low than on the exhaust side, thereby preventing abnormal combustion on the intake side of the combustion chamber. Knocking can be suppressed.

[Brief description of the drawings]

FIG. 1 is an explanatory diagram showing a state of an air-fuel mixture stagnating near an electrode of a spark plug at an ignition timing of a cylinder of a spark ignition type direct injection engine according to Embodiment 1 of the present invention.

FIG. 2 is an overall configuration diagram of an engine.

FIG. 3 is a perspective view showing an arrangement configuration of a piston crown, an intake port, a spark plug, and an injector.

FIG. 4 is a graph showing an example of a correspondence relationship between a fuel spray penetration force from an injector and a spray divergence angle.

FIG. 5 is a schematic configuration diagram of a fuel supply system.

FIG. 6 is a diagram illustrating an example of a control map in which an operation region in which the engine is in a stratified combustion state or a uniform combustion state is set.

7A is a top view showing the structure of a piston, FIG. 7B is a sectional view taken along line bb, and FIG. 7C is a sectional view taken along line cc.

FIG. 8 is an explanatory diagram showing a positional relationship among a concave portion of a piston crown surface, a tumble flow, and fuel spray, as viewed along a cylinder center line.

FIG. 9 is an explanatory diagram showing a geometric spray area, a spray center line, and a spray divergence angle of a fuel spray from an injector.

FIG. 10 is an explanatory diagram of (a) a spray divergence angle and (b) a spray penetration force.

FIG. 11 is a diagram corresponding to FIG. 1 at a fuel injection timing of a cylinder.

FIG. 12 is an explanatory diagram showing a change in the behavior of the spray with a change in the center line of the fuel spray.

FIG. 13 is a graph showing a change in a local air-fuel ratio near an electrode of a spark plug near a TDC in association with a change in an inclination angle of a fuel spray center line.

FIG. 14 is a graph showing a correspondence between a change in the fuel spray center line and a change in the amount of fuel adhering to the piston crown surface.

FIG. 15 shows an inclination angle of the fuel spray center line of 30 ° to 3 °.
It is a graph which shows the change of the local air-fuel ratio near the electrode of a spark plug when it is set to 6 degrees.

FIG. 16 is a graph showing a change in a local air-fuel ratio in accordance with a change in fuel injection timing.

FIG. 17 is a graph showing a correspondence relationship between a change in a spray spread angle and a change in a local air-fuel ratio near an electrode of a spark plug.

FIG. 18 is a diagram showing a state of the air-fuel mixture when the spray spread angle is approximately 20 °.

FIG. 19 is a diagram corresponding to FIG. 18 when the spray spread angle is approximately 60 °.

FIG. 20 is a diagram corresponding to FIG. 19 when a lemon-shaped recess is formed in the piston crown surface.

FIG. 21 is an explanatory diagram showing electrode positions of a spark plug in comparison with a tumble flow at a cylinder ignition timing.

FIG. 22 is a graph showing the relationship between the amount of protrusion of the spark plug electrode and the flow velocity of the tumble flow.

FIG. 23 is a graph showing the correspondence between the amount of protrusion of the spark plug electrode and the variation rate of combustion.

FIG. 24 is a graph in which the range of the protruding ratio of the spark plug electrode is set so as to correspond to the change in the fuel injection timing of the cylinder.

FIG. 25 is a diagram illustrating a result of CFD analysis in which a tumble flow and a fuel spray flow are balanced in a combustion chamber.

26A and 26B are diagrams showing the states of the air-fuel mixture when (a) the tumble flow and the spray penetration moderately collide and (b) when the collision is excessively severe, respectively.

FIG. 27 is a map diagram showing changes in the tumble flow velocity and the spray penetration force according to the engine rotation speed.

FIG. 28 is a map diagram showing a correspondence relationship between a spray penetration force and a fuel injection pressure.

FIG. 29 is a diagram corresponding to FIG. 27 according to a modification.

FIG. 30 is a schematic configuration diagram of an apparatus for measuring a tumble ratio.

FIG. 31 shows a change in the temperature state of the combustion chamber, (a) a change in the intake air temperature and the spray penetration force, and (b) a change in the fuel injection pressure and a change in the spray penetration force associated therewith. FIG.

FIG. 32 is a view corresponding to FIG. 1 in an intake stroke of a cylinder.

FIG. 33 is a graph showing a correspondence relationship between a fuel efficiency improvement rate, a fuel efficiency rate, an HC emission rate, and an engine load in a low rotation range of the engine in comparison with a conventional direct injection engine.

FIG. 34 is a diagram corresponding to FIG. 33 for an engine middle rotation range.

FIG. 35 is a diagram corresponding to FIG. 7 according to the second embodiment of the present invention.

FIG. 36 is a view corresponding to FIG. 35 according to another modification of the second embodiment.

FIG. 37 is a graph showing a change in a tumble ratio with respect to a crank angle for each shape of a concave portion of a piston crown surface.

FIG. 38 is a diagram corresponding to FIG. 35 according to another modification of the second embodiment.

FIG. 39 according to a further modification of the second embodiment.
FIG.

FIG. 40 is an explanatory diagram showing a shape of fuel spray from an injector according to Embodiment 3 of the present invention.

FIG. 41 is an explanatory diagram showing a relationship between fuel spray and a tumble flow in a third embodiment.

FIG. 42 is a diagram corresponding to FIG. 40 according to the fourth embodiment of the present invention.

FIG. 43 is a view corresponding to FIG. 41 according to the fourth embodiment.

FIG. 44 is a view corresponding to FIG. 43 according to a modification of the fourth embodiment.

FIG. 45 is an explanatory diagram of another embodiment in which the tumble flow velocity is adjusted by changing the valve timing.

FIG. 46 is an explanatory diagram of air-fuel mixture formation in a third conventional example.

FIG. 47 is a graph showing a trade-off relationship between an output improvement rate and a fuel efficiency improvement rate by a conventional direct injection engine.

[Explanation of symbols]

 Reference Signs List 1 spark ignition type direct injection engine 2 cylinder 5 piston 5a recess 6 combustion chamber 7 crankshaft 10 intake port (tumble flow generating means) 11 exhaust port 16 spark plug 18 injector (fuel injection valve) F fuel spray center line T tumble flow Tm Tumble forward flow Ts Tumble forward flow z Cylinder centerline θ Fuel spray spread angle

Continued on the front page (51) Int.Cl. ⁷ Identification symbol FI Theme coat II (reference) F02M 61/14 310 F02M 61/14 310A 310S 61/18 360 61/18 360J F02P 13/00 301 F02P 13/00 301A 302 302A (72) Inventor Hiroyuki Yamashita 3-1, Fuchi-machi, Shinchu, Aki-gun, Hiroshima Mazda Co., Ltd. Person Masakazu Matsumoto 3-1 Shinchi, Fuchu-cho, Aki-gun, Hiroshima Mazda Co., Ltd. (72) Inventor Fumihiko Saito 3-1 Shinchi, Fuchu-cho, Aki-gun, Hiroshima Pref. Mazda Co., Ltd. 3-1, Gunfuchu-cho Shinchi Mazda Co., Ltd. F-term (reference) 3G019 AA08 AA09 KA12 KA15 3G023 AA02 AB03 AC05 AD02 AD08 AD09 AG01 3G066 AA02 AA05 AA13 AB02 AD12 BA16 BA17 BA65 CC32 CC48 CD28 CD29 DC05 DC11 DC24

Claims (13)

[Claims] An ignition plug is provided at a ceiling of a combustion chamber facing a crown surface of a piston in a cylinder so as to protrude from near the center of the ceiling toward the combustion chamber. A fuel injection valve is disposed so as to inject fuel from the peripheral portion toward the center of the combustion chamber, and during stratified combustion operation, the fuel injected from the fuel injection valve is stratified around the electrode of the ignition plug. A spark ignition type direct injection engine, comprising: a tumble flow generating means capable of generating a tumble flow flowing between the electrode of the spark plug and the crown of the piston in the compression stroke of the cylinder toward the fuel injection valve; A concave portion is provided on the crown surface of the piston so as to follow the tumble flow, and the electrode of the ignition plug is provided after the fuel injection valve starts injection of fuel by the fuel injection valve in a compression stroke of a cylinder. And at a predetermined time before the ignition timing of the cylinder, the fuel injection valve is located at an intermediate portion between the tumble forward flow flowing along the recess and the tumble forward flow flowing along the combustion chamber ceiling. The direction in which the center line of the fuel spray intersects the bottom surface of the recess at the fuel injection timing of the cylinder during the stratified charge combustion operation is higher than the position where the center line of the cylinder intersects the bottom surface of the recess. A spark-ignition direct-injection engine, which is set so as to be away from the engine. 2. The fuel injector according to claim 1, wherein the electrode of the spark plug is located closer to the piston crown than the injection hole of the fuel injection valve when viewed from a direction perpendicular to the cylinder center line. Spark-ignition direct injection engine. 3. The ignition plug according to claim 1, wherein the electrode of the spark plug protrudes from the ceiling of the combustion chamber along the cylinder center line, and extends from the ceiling of the combustion chamber at the compression top dead center of the cylinder to the piston crown. When the distance on the cylinder center line is d, the distance e from the combustion chamber ceiling to the electrode of the ignition plug is e ≧
A spark-ignition direct-injection engine having a value satisfying a relation of 0.4d. 4. The fuel injection valve according to claim 1, wherein the fuel injection direction of the fuel injection valve is set such that the center line of the fuel spray passes between the electrode of the spark plug and the bottom surface of the concave portion of the piston crown surface. A spark ignition type direct injection engine characterized by the following features. 5. The fuel injection valve according to claim 1, wherein the fuel injection valve has a fuel injection direction having an inclination angle in a range of approximately 25 ° to approximately 40 ° with respect to an imaginary plane in which the center line of the fuel spray is orthogonal to the cylinder center line. A spark-ignition direct-injection engine, which is set to perform. 6. The fuel injection valve according to claim 1, wherein the shape of the fuel spray in the compression stroke of the cylinder is larger when viewed along the cylinder center line than when viewed along the axis of the crankshaft. A spark-ignition direct-injection engine, wherein an electrode of the spark plug is located outside a geometric fuel spray area from the fuel injection valve in a compression stroke of the cylinder. . 7. The fuel injection valve according to claim 1, wherein the fuel spray penetrating force in the compression stroke of the cylinder is greater on the side closer to the crown surface of the piston with respect to the center line of the fuel spray than on the opposite side. A spark-ignition direct-injection engine. 8. The fuel injection valve according to claim 1, wherein the fuel spray penetrating force in the compression stroke of the cylinder is larger on the side closer to the combustion chamber ceiling with respect to the center line of the fuel spray than on the opposite side. Wherein the electrode of the spark plug is located outside a geometric fuel spray area from the fuel injection valve during a compression stroke of the cylinder. 9. The fuel injection system according to claim 1, wherein the recess has a shape that is long in a direction in which the center line of the fuel spray extends when viewed along the cylinder center line. Spark-ignition type, characterized in that the width is wider than the width of the end near the injection valve, and that the side walls on both sides in the width direction project from the bottom surface to a predetermined height or more at a substantially central portion in the length direction of the recess. Direct injection engine. 10. The width of an end on a side farther from the fuel injection valve among ends on both sides in a longitudinal direction of the concave portion, which is substantially the same as a width of a substantially central portion in the longitudinal direction of the concave portion. A spark-ignition direct-injection engine. 11. The method according to claim 9, wherein, of the ends on both sides in the length direction of the recess, the width of the end farther from the fuel injection valve is wider than the substantially central portion in the length direction of the recess. Spark ignition type direct injection engine. 12. The squish area according to claim 1, wherein a squish area is formed in at least an outer peripheral portion on the exhaust side of the crown surface of the piston so as to extend along a ceiling of the opposed combustion chamber. Spark-ignition direct injection engine. 13. The fuel injection valve according to claim 1, wherein the fuel injection valve is located near the intake valve of the intake and exhaust valves, and the recess is formed by the fuel injection valve when viewed along the cylinder center line. A spark-ignition direct-injection engine, wherein the shape is long in a direction in which a center line extends, and the volume of the concave portion is smaller on a side closer to the fuel injector than on a cylinder center line than on the opposite side.