JP4793108B2 - In-cylinder direct injection internal combustion engine - Google Patents

Description

  The present invention relates to an in-cylinder direct injection internal combustion engine that is particularly intended for homogeneous combustion.

2. Description of the Related Art Conventionally, an in-cylinder direct injection internal combustion engine that directly injects fuel from a fuel injection valve into a cylinder in an internal combustion engine that performs spark ignition combustion is known. In such an in-cylinder direct injection internal combustion engine, as a combustion mode, there is homogeneous combustion that realizes stoichiometric operation (theoretical air-fuel ratio operation) by mainly performing fuel injection during the intake stroke (particularly the first half of the intake stroke). . At this time, the fuel injection start timing and the fuel injection period are determined from the engine speed and the target torque.
Japanese Patent No. 3161204

  However, the homogenization of the mixture of fuel and air in the combustion chamber is determined by various conditions in the combustion chamber, and it is desired to further promote the homogenization of the mixture in order to improve the stability of combustion.

  Accordingly, an object of the present invention is to provide an in-cylinder direct injection internal combustion engine that can promote homogenization of fuel distribution in the homogeneous combustion of an in-cylinder direct injection internal combustion engine.

  The present invention solves the above problems by the following means. In addition, in order to make an understanding easy, although the code | symbol corresponding to embodiment of this invention is attached | subjected, it is not limited to this.

  The present invention relates to a fuel injection valve (14) for directly injecting fuel into a combustion chamber defined by a piston of an internal combustion engine, prediction means (S103, S104) for predicting the gas flow strength in the combustion chamber, combustion Fuel injection control means (S106, S107) for controlling the fuel injection timing of the fuel injection valve according to the gas flow strength in the room.

The present invention relates to a fuel injection valve that directly injects fuel into a combustion chamber defined by a piston of an internal combustion engine, and a gas flow strength in the combustion chamber in the vicinity of the intake valve based on the engine speed and the intake valve lift amount. of predicting the gas flow strength, and prediction means to predict the gas flow strength by the piston behavior based on the engine speed, gas flow strength by the gas flow intensity and the piston behavior of the intake valves near predicted by said predicting means comprising a fuel injection control means to control the fuel injection timing of the fuel injection valve according to Sato, the.

(First embodiment)
Hereinafter, a first embodiment of the present invention will be described with reference to the drawings.

  FIG. 1 is a configuration diagram of an engine according to a first embodiment of the present invention, in which 1 is an engine and 2 is a cylinder block.

  In the cylinder block 2, a plurality of cylinders 3 are arranged in series, and a cylinder head 4 is fixed so as to cover the upper surface thereof. A piston 5 is slidably fitted in the cylinder 3, and a combustion chamber 6 is formed above the piston 5.

  The cylinder head 4 is formed with an intake port 7 and an exhaust port 8 that open to the combustion chamber 6, and an intake valve 9 and an exhaust valve 10 that open and close the ports 7 and 8. The intake valve 9 is driven to open and close by an intake valve cam 11, and the exhaust valve 10 is driven to open and close by an exhaust valve cam 12.

  In the cylinder head 4, a spark plug 13 is disposed opposite to the substantially center of the cylinder 3 so that the tip of the cylinder head 4 faces the combustion chamber 6.

  Further, the electromagnetic fuel injection valve 14 is disposed immediately below the intake port 7 in a posture in which the central axis thereof is inclined obliquely downward, and the tip thereof is so as to inject fuel directly into the combustion chamber 6. The part faces the combustion chamber 6.

  Further, the engine 1 includes a variable valve mechanism 16 capable of variably controlling the operating angle and the phase of the operating angle of the intake valve 9, and a variable compression ratio mechanism capable of changing the compression ratio by changing the stroke characteristics of the piston 5. 17.

  The spark plug 13, the fuel injection valve 14, the variable valve mechanism 16 and the variable compression ratio mechanism 17 are controlled by an engine control unit 15 (hereinafter referred to as “ECU 15”). The ECU 15 calculates a required fuel injection amount and injection timing according to the engine speed, load, etc., generates a fuel pressure required for a fuel pump (not shown), outputs an injection valve drive signal, and controls fuel injection. Then, an ignition signal is output according to the engine speed, load, etc., and an ignition coil (not shown) is driven and controlled, and the ignition plug 13 is discharged to control ignition.

  The configuration and operation of the variable compression ratio mechanism 17 will be described with reference to FIG.

  The crankshaft 21 includes a plurality of journals 21a and a crankpin 21b. The journal 21a is rotatably supported by a main bearing (not shown) of the cylinder block 2. The crank pin 21b is eccentric from the journal 21a by a predetermined amount, and a central portion of the lower link 22 serving as the second link is rotatably connected thereto.

  The lower link 22 is configured to be split into two left and right members, and a crank pin 21b is fitted in a substantially central connecting hole.

  The upper link 23 serving as the first link has a lower end side rotatably connected to one end of the lower link 22 by a first connecting pin 24, and an upper end side rotatably connected to the piston 5 by a piston pin 25. The piston 5 receives the combustion pressure and reciprocates in the cylinder 3 of the cylinder block 2.

  The control link 26 serving as the third link is pivotally connected to the other end of the lower link 22 by the second connecting pin 27 at the upper end side, and the cylinder block 2 whose lower end side is part of the engine body via the control shaft 28. It is connected with the lower part of this so that rotation is possible.

  An eccentric sleeve bearing (not shown) is provided on the lower end side of the control link 26, and an eccentric shaft portion 29 of the control shaft 28 is rotatably inserted into the eccentric sleeve bearing. The control shaft 28 is rotated by an actuator 30 with a motor.

  When the control shaft 28 is rotated by the motor-equipped actuator 30, the center position of the eccentric shaft portion 29 changes, and the relative position with respect to the engine body changes. Thereby, the rocking | fluctuation support position of the lower end of the control link 26 changes.

  When the swing support position of the control link 26 is changed, the stroke of the piston 5 is changed, and the position of the piston 5 at the piston top dead center is increased or decreased (see FIG. 3). This makes it possible to change the engine compression ratio.

  At the time of high compression ratio control, as shown in FIG. 3 (a), the control link 26 is lowered downward by the eccentric shaft portion 29, whereby the lower link 22 moves clockwise, and the first connecting pin 24 is moved. Since it is raised, the position of the top dead center of the piston 5 rises.

  At the time of low compression ratio control, as shown in FIG. 3B, the control link 26 is raised upward by the eccentric shaft portion 29, whereby the lower link 22 moves counterclockwise, and the second connecting pin 24 Is lowered, the position of the top dead center of the piston 5 is lowered.

  3 (a) and 3 (b) representatively show a high compression ratio state and a low compression ratio state, the variable compression ratio mechanism 17 continuously sets the compression ratio between them. Can be changed.

  FIG. 4 shows piston characteristics in the variable compression ratio mechanism. As shown in FIG. 4, the behavior of the piston 5 in the variable compression ratio mechanism 17 using the multi-link type piston / crank mechanism as described above is compared with the case of the normal link mechanism. Is about 80 degrees after the top dead center, whereas the time when the piston speed of the variable compression ratio mechanism 17 is maximum is about 100 to 120 degrees after the top dead center on the slightly retarded side. .

  Based on FIG. 5 which shows the structure of the variable valve mechanism 16, the structure and operation | movement are demonstrated.

  The variable valve mechanism 16 is a variable valve mechanism that can continuously change the intake valve lift characteristics, and continuously changes the valve lift amount and the operating angle of the intake valve 9 by changing the rotation angle of the control shaft 41. A first variable valve mechanism 42 that can be changed, and a second variable valve mechanism 43 that can continuously change the center phase of the operating angle of the intake valve 9 to the advance side and the retard side. .

  The intake camshaft 44 receives rotational power from the crankshaft 21 via a pulley or sprocket 53 and rotates about its own axis in conjunction with the rotation of the crankshaft 21.

  A swing cam 46 that opens and closes the intake valve 10 is rotatably fitted and supported on the intake cam shaft 44 above the valve lifter 45 of the intake valve 9.

  The first variable valve mechanism 42 includes a circular eccentric cam 47 that is eccentrically fixed to or integrally formed with the intake camshaft 44, a control shaft 41 that extends parallel to the intake camshaft 44 in the cylinder row direction, and the control shaft 41. A circular control cam 48 that is eccentrically fixed or integrally formed, a rocker arm 49 that is swingably fitted to the control cam 48, and one end of the rocker arm 49 and the eccentric cam 47 that cooperate with each other. A ring-shaped first link 50; a rod-shaped second link 51 that links the other end of the rocker arm 49 and the tip of the swing cam 46; a first actuator 52 that changes the rotation angle of the control shaft 41; have.

  One end of the first link 50 is attached to the eccentric cam 47 so as to be rotatably fitted.

  When the intake camshaft 44 rotates in conjunction with the crankshaft 21, one end of the first link 50 fitted to the eccentric cam 47 is rotationally displaced with respect to the axial center of the intake camshaft 44, so that the first link 50 is As a whole, the translation cam operates substantially, and the swing cam 46 swings within a predetermined swing angle range via the rocker arm 49 and the second link 51 linked to the first link 50.

  The intake valve 9 opens and closes when the swing cam 46 pushes down the valve lifter 45.

  When the rotation angle of the control shaft 41 is changed by the first actuator 52, the center of the control cam 48 that is the swing center of the rocker arm 49 is rotationally displaced with respect to the axis of the control shaft 41, and the links 50 and 51 are moved. As a result, the initial posture of the swing cam 46 (the center phase of the swing range) changes.

  As a result, the operating angle of the intake valve 10 and the valve lift amount continuously change while the phase of the operating center angle of the intake valve 9 with respect to the crank angle remains substantially constant.

  The second variable valve mechanism 43 is housed in an external rotating body 53 including a sprocket (or cam pulley) that rotates in conjunction with the crankshaft 3 and the external rotating body 53, and is integrated with the intake camshaft 44. An internal rotating body (not shown) that rotates, and by rotating both the rotating bodies relative to each other by hydraulic pressure, the phase of the operation center angle (intake center angle) of the intake valve 9 with respect to the crank angle, that is, the intake valve 9 The valve timing is advanced or retarded.

  FIG. 6 shows lift amounts of the intake valve 9 and the exhaust valve 10 in the present invention. As shown in FIG. 6, the lift amount of the intake valve 9 including the variable valve mechanism 16 is changed depending on the operating state of the engine 1. That is, when the engine 1 is under a low load, the lift amount is decreased as shown in the characteristic 62, and the valve closing timing is advanced more than the bottom dead center. Thereby, since the intake air amount can be limited in the intake valve 9, the throttle opening can be opened, and the pump loss can be reduced.

  The variable valve mechanism 16 increases the intake valve lift amount as the load increases, and retards the valve closing timing. For example, in FIG. 6, when the engine 1 is medium load, the lift amount is larger than the low load characteristic 62 as shown by the characteristic 63, the valve closing timing is retarded, and when the engine 1 is full load, the characteristic 64 shows. In addition, the lift amount is set to be larger than that of the medium load so that the valve closing timing is retarded.

  On the other hand, the lift amount of the exhaust valve 10 that does not have the variable valve mechanism is uniquely determined with respect to the crank angle regardless of the load, as indicated by the characteristic 61.

  FIG. 7 is a graph showing the relationship between the gas flow strength in the combustion chamber 6 of the engine 1 and the fuel injection period by the fuel injection valve 14. Based on FIG. 7, the concept of the present invention will be described.

  In the graph of FIG. 7, the horizontal axis is constituted by the crank angle, and the vertical axis is constituted by the gas flow strength. The gas flow strength is expressed by gas flow rate × gas flow rate.

  In the engine 1 having the variable valve mechanism 16 and the variable compression ratio mechanism 17, the maximum factor that governs the gas flow in the combustion chamber 6 is the behavior of the intake valve 9 or the piston 5. That is, as shown in FIG. 8, in the combustion chamber 6, the gas flow A in the vicinity of the intake valve 9 due to the air flowing into the combustion chamber 6 from the intake port 7 when the intake valve 9 is open, and the inside of the cylinder 3. Two main gas flows of gas flow B are generated by the piston behavior generated in the vicinity of the crown surface of the piston 5 that moves up and down. The gas flow A in the vicinity of the intake valve 9 is greater in the gas flow strength in the vicinity of the spark plug 13 than in the vicinity of the fuel injection valve 14.

  The gist of the present invention is to promote homogenization of the fuel distribution in the combustion chamber 6 by controlling the fuel injection timing in accordance with the gas flow strength.

  FIG. 7A is a graph showing the normal valve lift amount, that is, the gas flow rate strength and the fuel injection period when the variable valve mechanism 16 is not operating. As shown by the characteristic 64 in FIG. 6, since the lift amount of the intake valve 9 is set to the maximum at full load or the like, in the combustion chamber 6, the gas flow in the vicinity of the intake valve is more than the gas flow due to the piston behavior. Become dominant. Therefore, the fuel injection period is set to a time when the gas flow strength near the intake valve becomes large, and the approximate center of the fuel injection period is set near the maximum of the gas flow strength near the intake valve.

  FIG. 7B is a graph showing the gas flow rate strength and the fuel injection period when the lift amount of the intake valve is a medium lift and the opening / closing valve timing is advanced. As shown by the characteristic 63 in FIG. 6, in a high load region where the load is lower than the full load, an intermediate load, or the like, the lift amount of the intake valve is set smaller than the normal valve lift amount, and the opening / closing valve timing is advanced. Therefore, the peak of the gas flow strength flowing from the intake valve 9 also moves to the advance side. Even in this case, in the combustion chamber 6, the gas flow in the vicinity of the intake valve is more dominant than the gas flow generated by the piston behavior. Therefore, in a high load flow region or a medium load where the load is lower than the full load, The fuel injection period is set to a time when the gas flow strength near the intake valve becomes large, and the approximate center of the fuel injection period is set near the maximum of the gas flow strength near the intake valve.

  FIG. 7C is a graph showing the gas flow rate strength and the fuel injection period when the lift amount of the intake valve is a small lift amount and the opening / closing valve timing is advanced. As shown by the characteristic 62 in FIG. 6, at a low load or the like, the lift amount of the intake valve 9 is set to be smaller than that at the middle load, and the on-off valve timing is advanced. Therefore, the gas flow strength sucked from the intake valve 9 is reduced, and the gas flow generated by the piston behavior becomes dominant. Therefore, at a low load or the like, the fuel injection period is set to the time when the gas flow strength due to the piston behavior becomes large, and the approximate center of the fuel injection period is set near the maximum of the gas flow strength due to the piston behavior.

  FIG. 9 and FIG. 10 are a control flowchart and a map used for the control in this embodiment.

  FIG. 9 is a control flow executed by the ECU 15 at regular intervals. Hereinafter, based on FIG. 9, the control in this embodiment is demonstrated.

  First, in step S101, the engine speed and load are read. In step S102, the intake valve opening / closing timing and the lift amount are read. The intake valve opening / closing timing and the lift amount are determined by the variable valve mechanism 16 according to the rotation and load.

  In step S103, the gas flow strength near the intake valve is predicted. FIG. 10A is a map used for predicting the gas flow strength near the intake valve. The gas flow strength in the vicinity of the intake valve can be predicted by the engine speed and the intake valve lift amount. The map in FIG. 10 (a) is composed of the rotation speed on the horizontal axis and the intake valve lift amount on the vertical axis. ing. As shown in FIG. 10A, the gas flow strength in the vicinity of the intake valve increases as the engine speed increases and as the intake valve lift amount increases.

  Next, in step S104, the gas flow strength due to the piston behavior is predicted. FIG. 10B is a map used for prediction of gas flow strength based on piston behavior. The gas flow strength due to the piston behavior can be predicted from the engine rotational speed, and the map of FIG. 10 (b) is composed of the engine rotational speed on the horizontal axis and the gas flow strength due to the piston behavior on the vertical axis. . As shown in FIG. 10B, the gas flow strength due to the piston behavior increases as the engine speed increases.

  Subsequently, in step S105, the gas flow strength in the vicinity of the intake valve predicted in step S103 is compared with the gas flow strength in the vicinity of the intake valve predicted in step S104. Determine whether it is greater than the flow strength. When the gas flow strength in the vicinity of the intake valve is larger than the gas flow strength due to the piston behavior, the process proceeds to step S106.

  In step S106, the center of the fuel injection period is set near the maximum gas flow strength near the intake valve, that is, near the maximum lift of the intake valve, and this flow ends.

  On the other hand, if the gas flow strength in the vicinity of the intake valve is smaller than the gas flow strength due to the piston behavior in step S105, the process proceeds to step S107, where the gas flow strength due to the piston behavior is near the center of the fuel injection period. That is, it is set near the time when the piston speed becomes maximum, and the process proceeds to step S108.

  In step S108, the fuel injection pressure of the fuel injection valve 14 is set higher than when the gas flow strength in the vicinity of the intake valve is larger than the gas flow strength due to the piston behavior, and this flow ends.

  As described above, according to the present embodiment, since the fuel injection timing is determined according to the most dominant gas flow strength in the combustion chamber 6, the fuel is supplied to the strong gas flow field, and the fuel is vaporized. Promoted to homogenize the fuel distribution.

  Further, since the center of the fuel injection timing is set at the time when the most dominant gas flow strength in the combustion chamber 6 is maximized, fuel is supplied to the strongest gas flow field, and fuel vaporization is further promoted. The fuel distribution is homogenized.

  In particular, the gas flow strength in the vicinity of the intake valve and the gas flow strength due to piston behavior are predicted and compared, and fuel is injected according to the gas flow strength of whichever is greater, so the fuel flows into a strong gas flow field. Is supplied, fuel vaporization is promoted, and the mixture of fuel and air is made more homogeneous.

  Also, when the gas flow strength near the intake valve is smaller than the gas flow strength due to the piston behavior, the fuel injection pressure is higher than when the gas flow strength near the intake valve is greater than the gas flow strength due to the piston behavior. Since the gas flow strength in the vicinity of the intake valve is small and it is difficult to ensure uniform fuel distribution, the penetration force of the spray is transmitted to the gas flow in the combustion chamber 6 and the combustion chamber 6 The gas flow can be strengthened to make the fuel distribution uniform.

  Furthermore, the time when the gas flow strength due to the piston behavior becomes maximum, such as the variable compression ratio mechanism 17, is retarded compared to the normal piston / crank mechanism, and the gas flow strength near the intake valve becomes maximum. Since the present invention is applied to an internal combustion engine having a multi-link type piston / crank mechanism in which the gas flow strength due to piston behavior is maximized, the time when the gas flow strength is maximized is thus shifted. However, it is possible to determine the fuel injection timing suitable for homogenizing the fuel distribution, and to promote the homogenization of the mixture of fuel and air.

  Also, the multi-link type piston / crank mechanism has a longer stay time near the top dead center of the piston and the gas flow strength is stronger than the normal piston / crank mechanism, so the internal combustion engine equipped with the multi-link type piston / crank mechanism. By applying the present invention to the engine, the effect of the present invention for achieving homogenous fuel distribution using gas flow strength can be further obtained.

In addition, the higher the engine compression ratio, the longer the stay time near the top dead center of the piston. Therefore, by applying the present invention to an internal combustion engine equipped with a variable compression ratio mechanism, the fuel distribution is utilized using the gas flow strength. The effect of the present invention for achieving homogenization can be further obtained.
(Second Embodiment)
Hereinafter, a second embodiment of the present invention will be described with reference to FIG.

  The present embodiment is different from the first embodiment in the method of determining the fuel injection timing and the fuel injection pressure in the ECU 15. The engine configuration is the same as in the first embodiment. FIG. 11 is a fuel injection timing map for obtaining the fuel injection timing in the second embodiment.

  In the fuel injection timing map of FIG. 11, the horizontal axis is composed of the engine load and the vertical axis is composed of the fuel injection timing, and the fuel injection timing (period) is obtained based on the engine load. In this fuel injection timing map, a predetermined load in which the gas flow strength in the vicinity of the intake valve and the gas flow strength due to the piston behavior are interchanged is obtained in advance, and in a load region lower than the predetermined load, the gas due to the piston behavior is obtained. Assuming that the flow strength is greater than the gas flow strength in the vicinity of the intake valve, the map is set so that the fuel injection timing matches the timing when the gas flow strength due to the piston behavior is maximized.

  Further, in a load region higher than a predetermined load, it is assumed that the gas flow strength near the intake valve is larger than the gas flow strength due to the piston behavior, and the fuel injection period is reached when the gas flow strength near the intake valve becomes maximum. The map is set so that the approximate centers of

  FIG. 12 is a graph showing the relationship between the engine speed and a predetermined load in which the gas flow strength near the intake valve used in the fuel injection timing map of FIG.

  The gas flow strength in the vicinity of the intake valve is determined by the engine speed and the lift amount of the intake valve 9. The higher the engine speed and the greater the lift amount of the intake valve 9, the greater the gas flow strength in the vicinity of the intake valve. Become. Since the lift amount of the intake valve 9 is set to increase as the engine load increases, the predetermined load is set to decrease as the engine speed increases as shown in FIG. 12, and the ECU 15 performs a plurality of fuel injections according to the engine speed. Has a map.

  The fuel injection pressure is determined based on the fuel injection pressure map shown in FIG. In this fuel injection pressure map, the horizontal axis is composed of the engine speed and the vertical axis is composed of the load.

  In a region where the load is lower than the boundary line 72, the lift amount change control of the intake valve 9 is performed in order to reduce the pump loss, and the lift amount is reduced and the valve closing timing is advanced from the bottom dead center. The gas flow in 6 is weak and it is difficult to ensure combustion stability. Therefore, the fuel injection pressure is set higher as the engine speed and load are lower.

  Specifically, in the region sandwiched between the boundary line 72 and the boundary line 73 that is exactly the same as the predetermined load graph shown in FIG. 12, the gas flow strength in the vicinity of the intake valve is considered to be dominant. Therefore, the fuel injection pressure is set to a medium level. Further, in the low load region from the boundary line 73, the gas flow due to the piston behavior becomes dominant and the gas flow in the vicinity of the intake valve is considered weak, so the fuel injection pressure is set to the maximum value that can be set.

  On the other hand, in the high load region from the boundary line 72, the lift amount change control of the intake valve 9 is not performed, and the lift amount is large and the gas flow in the combustion chamber 6 is strong. Therefore, since it is easy to ensure the combustion stability, the fuel injection pressure is lowered at the same engine speed as compared with the low load region from the boundary line 72 in order to reduce the mechanical loss of driving the fuel pump. However, when the engine speed is higher in the high load region than the boundary line 72 and the fuel injection period cannot be taken sufficiently, the fuel injection pressure is set to the maximum value that can be set and the fuel injection period is set to be short. .

  In FIG. 13, the fuel injection pressure is switched in three stages of low, medium, and high, but may be switched continuously.

  According to this embodiment, since the fuel injection timing control is switched by a predetermined load, it is not necessary to predict and compare the gas flow strength near the intake valve and the gas flow strength due to the piston behavior, and the control can be switched quickly. , Drivability is improved.

  In addition, in a low load range where the lift amount of the intake valve 9 is small and the gas flow is weak and combustion stability cannot be secured, the fuel injection pressure is set higher as the engine speed and load are lower. Even when the fuel distribution is small and it is difficult to ensure homogenization of the fuel distribution, the penetration force of the spray is transmitted to the gas flow in the combustion chamber 6 to strengthen the gas flow in the combustion chamber 6 and Homogenization can be achieved.

  Further, in the high load region where the lift amount of the intake valve 9 is large and the gas flow in the combustion chamber 6 is strong, the fuel injection pressure is set lower than in the low load region, so that the mechanical loss of the fuel pump drive can be reduced.

In addition, since the predetermined load is reduced as the engine speed increases, the range in which the fuel injection pressure is set high can be minimized.
(Third embodiment)
Hereinafter, a third embodiment of the present invention will be described with reference to FIGS. 14 and 15.

  This embodiment is particularly different from the first and second embodiments in the engine configuration. In the first embodiment, the present invention is applied to the engine 1 having the variable compression ratio mechanism 17, but the present invention can also be applied to a multi-link type piston / crank mechanism internal combustion engine having no variable compression ratio mechanism. .

  FIG. 14 is a diagram of a piston in the third embodiment. 14 (a) is a perspective view of the short skirt piston 71, FIG. 14 (b) is a cross-sectional view taken along the line BB of FIG. 14 (a), and FIG. 14 (c) is a cross-sectional view of FIG. It is C sectional drawing.

  As shown in FIG. 14, the short skirt piston 71 has a significantly shortened piston skirt.

  If such a short skirt piston 71 is used, the counterweight 62 can pass through the side of the piston pin 25 as shown in FIG. Therefore, by setting the upper link 23 to the minimum length and bringing the bottom dead center position of the short skirt piston 71 closest to the crankshaft 21, the piston stroke can be expanded accordingly. In order to make such a configuration, the strength of the piston skirt is an issue, but as shown in FIG. By making the upper link 23 substantially upright at the top dead center position, the lateral load (thrust load) applied to the short skirt piston 71 can be reduced. Thereby, the strength of the piston skirt is ensured.

  As shown in FIG. 4, the piston behavior of the multi-link piston-crank mechanism internal combustion engine as described above is retarded when the piston speed becomes maximum as compared with the piston-crank mechanism of the normal link. The

  The time when the gas flow strength due to the piston behavior becomes the maximum as in this embodiment is retarded compared to the normal piston / crank mechanism, and the time when the gas flow strength near the intake valve becomes the maximum depends on the piston behavior. Since the present invention is applied to an internal combustion engine having a multi-link type piston / crank mechanism in which the time at which the gas flow strength becomes maximum is shifted, the fuel distribution can be obtained even when the time at which the gas flow strength is maximized is shifted. The fuel injection timing suitable for the homogenization of the fuel can be determined, and the homogenization of the mixture of fuel and air can be promoted.

  In addition, the multi-link piston / crank mechanism has a longer stay time near the top dead center of the piston and a stronger gas flow strength than a normal piston / crank mechanism. By applying the above, it is possible to further obtain the effect of the present invention for homogenizing the fuel distribution using the gas flow strength.

  The present invention is not limited to the embodiment described above, and various modifications and changes can be made within the scope of the technical idea, and it is obvious that these are equivalent to the present invention.

  For example, in the embodiment of the present invention, the present invention is applied to an engine in which the gas flow strength in the vicinity of the intake valve is changed by the variable valve mechanism, but the variable valve mechanism is not provided, and the swirl control valve is installed in the intake port. You may apply this invention to the engine with which it prepares. In this case, the flow strength of the intake air flowing into the combustion chamber changes depending on the degree of opening and closing of the swirl control valve, and the gas flow strength in the vicinity of the intake valve can be predicted using the map of FIG.

It is a block diagram of the engine in 1st Embodiment. It is a block diagram of the variable compression ratio mechanism in 1st Embodiment. It is the schematic which showed the link position of the variable compression ratio mechanism in 1st Embodiment. It is a piston characteristic view of the variable compression ratio mechanism in the first embodiment. It is a block diagram of the variable valve mechanism in 1st Embodiment. It is a graph which shows the lift amount of the intake valve and exhaust valve in 1st Embodiment. It is the graph which showed the relationship between the gas flow strength in the combustion chamber in 1st Embodiment, and the fuel injection period by a fuel injection valve. It is the schematic which shows the gas flow by the piston behavior near the gas flow in the combustion chamber vicinity in 1st Embodiment. It is a control flowchart in a 1st embodiment. It is also a map used for the control. It is a fuel-injection time map for calculating | requiring the fuel-injection time in 2nd Embodiment. It is a graph which shows the relationship between the predetermined load and engine speed in 2nd Embodiment. 3 is a fuel injection pressure map for determining fuel injection pressure from engine speed and load. It is a figure of the short skirt piston in 3rd Embodiment. It is a block diagram of the multilink type piston-crank mechanism in 3rd Embodiment. It is a map which calculates | requires the gas flow strength of the intake valve vicinity in an engine provided with a swirl control valve.

Explanation of symbols

1 Engine 2 Cylinder block 3 Cylinder 4 Cylinder head 5 Piston 6 Combustion chamber 7 Intake port 8 Exhaust port 9 Intake valve 10 Exhaust valve 11 Intake valve cam 12 Exhaust valve cam 13 Spark plug 14 Fuel injection valve 15 ECU
16 Variable valve mechanism 17 Variable compression ratio mechanism 21 Crankshaft 21a Journal 21b Crank pin 22 Lower link 23 Upper link 24 First link pin 25 Piston pin 26 Control link 27 Second link pin 28 Control shaft 29 Eccentric shaft part 30 With motor Actuator 41 Control shaft 42 First variable valve mechanism 43 Second variable valve mechanism 44 Intake camshaft 45 Valve lifter 46 Oscillating cam 47 Eccentric cam 48 Control cam 49 Rocker arm 50 First link 51 Second link 52 First actuator 53 External rotating body 71 Short skirt piston

Claims (15)

A fuel injection valve for directly injecting fuel into the combustion chamber defined by the piston of the internal combustion engine;
Predicting the gas flow strength near the intake valve based on the engine speed and the intake valve lift as the gas flow strength in the combustion chamber, and predicting the gas flow strength due to piston behavior based on the engine speed Means,
Fuel injection control means for controlling the fuel injection timing of the fuel injection valve according to the gas flow strength in the vicinity of the intake valve predicted by the prediction means and the gas flow strength due to piston behavior ;
An in-cylinder direct injection internal combustion engine comprising: The fuel according to the prior SL fuel injection control means, the predicting means compares the gas flow strength by gas flow intensity and the piston behavior of the intake valves near the predicted, or strong and the determined towards the gas flow strength of The in-cylinder direct injection internal combustion engine according to claim 1, wherein the fuel injection timing of the injection valve is controlled. The prediction means, the higher the institutional rpm, also cylinder direct injection internal combustion engine according to claim 2 determines that the strong gas flow strength in the vicinity of the intake valve larger the intake valve lift amount. The prediction means, the higher the institutional rpm, cylinder direct injection internal combustion engine according to claim 2 or 3 determines that the gas flow strength by the piston behavior is strong.   When the fuel injection timing of the fuel injection valve is set according to the gas flow strength due to the piston behavior, the fuel injection control means is more than when it is set according to the gas flow strength near the intake valve. The in-cylinder direct injection internal combustion engine according to any one of claims 2 to 4, wherein increases a fuel injection pressure of the fuel injection valve. The fuel injection control means determines the fuel injection timing by the fuel injection valve,
When the load of the internal combustion engine is lower than the predetermined value, set according to the gas flow strength by the piston behavior,
2. The direct injection type internal combustion engine according to claim 1, wherein when the load on the internal combustion engine is higher than a predetermined value, the internal combustion engine is set according to the gas flow strength in the vicinity of the intake valve.   When the fuel injection control means controls the fuel injection pressure based on the engine speed and the engine load, and the load on the internal combustion engine is lower than the predetermined value, the load on the internal combustion engine is higher than the predetermined value. The in-cylinder direct injection internal combustion engine according to claim 6, wherein the fuel injection pressure is increased as compared with the direct injection internal combustion engine.   The direct injection internal combustion engine according to claim 6 or 7, wherein the fuel injection control means sets the predetermined value to be smaller as the engine speed increases. A first link coupled to the piston via a first coupling pin;
A second link that is rotatably mounted on a crankpin of the crankshaft and connected to the first link via a second connecting pin;
A third link that is connected to the second link via a third connection pin and swings about a fulcrum provided on the cylinder block;
The in-cylinder direct injection internal combustion engine according to any one of claims 1 to 8, further comprising a multi-link type piston / crank mechanism for fastening the crank and the piston.   The in-cylinder direct injection type according to claim 9, wherein the multi-link type piston / crank mechanism is a variable compression ratio mechanism capable of changing a compression ratio of an internal combustion engine by changing a position of a fulcrum provided in the cylinder block. Internal combustion engine.   The in-cylinder direct injection internal combustion engine according to any one of claims 1 to 10, further comprising a variable valve mechanism that makes the operating angle of the intake valve and the phase of the operating angle variable.   The in-cylinder according to any one of claims 1 to 11, wherein the fuel injection control means controls a substantially center of a fuel injection timing of the fuel injection valve to a timing at which a gas flow strength in the combustion chamber becomes maximum. Direct injection internal combustion engine.   The time when the gas flow strength becomes the maximum is the time when the intake valve lift amount becomes the maximum in the gas flow strength near the intake valve, and the time when the piston speed becomes the maximum in the gas flow strength due to the piston behavior The in-cylinder direct injection internal combustion engine according to claim 12, wherein A fuel injection valve for directly injecting fuel into the combustion chamber defined by the piston of the internal combustion engine;
A swirl generating means provided in the intake port for generating an intake swirl in the combustion chamber;
As the gas flow strength in the combustion chamber, the gas flow strength near the intake valve is predicted based on the engine speed and the opening of the swirl generating means, and the gas flow strength due to the piston behavior is calculated based on the engine speed. Prediction means to predict;
Fuel injection control means for controlling the fuel injection timing of the fuel injection valve according to the gas flow strength in the vicinity of the intake valve predicted by the prediction means and the gas flow strength due to piston behavior;
An in- cylinder direct injection internal combustion engine comprising: The prediction means, the higher the institutional rpm, also said with the opening is small enough intake valves gas flow strength in the vicinity of the swirl generating means is determined to strong, the higher the engine speed, by the piston behavior The direct injection type internal combustion engine according to claim 14, wherein the gas flow strength is determined to be strong .