JP2000045781A - Cylinder injection type engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To realize the super-lean burn, to widen a range of an operable air/ fuel ratio to improve the fuel consumption and to reduce the exhaust by enabling the stratified burining even by a flat piston in a cylinder injection engine. SOLUTION: In a cylinder fuel injection-type engine comprising an injector 1 for directly injecting the fuel into a combustion chamber 13 formed between a cylinder head and a piston and an ignition plug 2 mounted on the cylinder head, a swirl forming means for forming the air current of lateral swirl revolving around a cylinder axis in a combustion chamber 13 in the intake stroke in a lean-burn combustion mode with the distribution of the swirl strength in the cylinder axial direction satisfying that the swirl strength at the ignition plug 2 side is higher than that at the piston 6 side.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an in-cylinder injection engine for directly injecting fuel into a cylinder (cylinder) of an engine having an ignition plug, and more particularly, to a super-lean burn in a low load operation range. The present invention relates to a stratified combustion technique for realizing ultra-lean mixture combustion.

[0002]

2. Description of the Related Art In recent years, in a gasoline engine having a spark plug, a direct injection engine for directly injecting fuel into a combustion chamber has been put to practical use.

[0003] In a direct injection type engine, when homogeneous combustion (uniform combustion) of an air-fuel mixture is intended, in order to homogenize the air-fuel mixture, the air taken into a combustion chamber (in a cylinder) is tumbled ( A method of forming a vertical swirl is known.

[0004] In order to realize lean burn in a low-load operation range, there is known a method of performing so-called stratified combustion by locally enriching only the air-fuel mixture around the spark plug.

For example, in a direct injection engine disclosed in Japanese Patent Application Laid-Open No. 8-35429, a cavity (recess) is provided on an upper surface of a piston, and fuel spray is confined in the cavity, thereby mixing fuel around the spark plug. The ignitability is ensured even when the mixture is lean and the entire cylinder is lean.

FIG. 2 shows a combustion mode using a conventionally provided piston with a cavity. In FIG. 2, 1 is an injector (fuel injection valve), 2 is a spark plug, 3 is fuel spray, 3 'is a mixture of fuel and air, 6 is a piston, 6a
Is a cavity, 7 is an engine cylinder (one of a plurality of cylinders is taken here), 13 is a combustion chamber, 14 is a cylinder head, 16 is an intake valve, and 17 is an exhaust valve.

FIG. 2 (a) shows a fuel injection state in a lean burn mode executed during a low load operation, in which the fuel is directed toward the cavity 6a before reaching the top dead center of the piston 6 during the compression stroke. By the injection, the fuel spray 3 is confined in the cavity 6a, and the fuel can be collected around the ignition plug 2. FIG. 2B shows a fuel injection state in a stoichiometric (ideal air-fuel ratio 14.7) combustion mode executed at the time of medium and high load operation, and the air and fuel are injected by injecting the fuel during the intake stroke. The mixing time is increased, and the mixing of fuel and air is promoted by the tumble (vertical swirl) airflow.

[0008] When the cavity is provided in the piston as described above, the surface area of the piston increases, and heat loss increases.
This is a cause of hindering improvement in fuel efficiency. Further, when a homogeneous mixture is formed in the cylinder (homogeneous combustion mode) as in the case of a full-open operation (high load), an excessive portion of fuel is generated in the cavity, and the tumble air flow is disturbed. It is conceivable to prevent homogenization of the mixture. If the mixture is not homogenized during high-load operation, it causes soot and output reduction.

FIG. 3 shows a state of air flow in a cylinder when a piston having a flat top surface is used. Since no cavities are provided, a tumble airflow is formed in the cylinder, and the mixing of air and fuel is promoted. Therefore, it is considered that fuel consumption and output are advantageous as compared with the piston with cavities during high load operation.

FIG. 4 shows the engine torque of a piston with cavities and a flat piston. A flat piston can improve torque compared to a piston with cavities.

[0011]

As described above, the flat piston is more advantageous than the piston with the cavity in terms of the output and the fuel consumption during homogeneous combustion, and has a smaller surface area than the piston with the cavity. It has the advantage of less loss.

However, it is difficult to collect fuel around the spark plug and the lean air-fuel ratio cannot be increased, so that the pumping loss is not reduced. Therefore, the effect of reducing fuel consumption during low load operation is small. If stratified combustion can be realized with a flat piston, fuel efficiency can be greatly reduced and a large amount of EGR (combustion gas) can be applied.
NOx can be reduced.

The present invention has been made in view of the above points, and an object of the present invention is to realize a super-lean burn by enabling stratified combustion even with a flat piston, and to achieve a wide range of operable air-fuel ratio in a cylinder. It is to realize an injection engine.

[0014]

In order to achieve the above object, the present invention basically provides an injector for directly injecting fuel into a combustion chamber formed between a cylinder head and a piston, and a cylinder head. In a direct injection type engine equipped with a spark plug arranged in a lean air-fuel mixture combustion mode, a swirl air flow swirling around a cylinder axis is introduced into a combustion chamber during an intake stroke in a lean mixture combustion mode, and a swirl strength on a spark plug side. Is provided with a swirl forming means for forming a swirl intensity distribution in the cylinder axis direction such that the swirl distribution is larger than the piston side.

According to the above configuration, during low load operation (lean burn), prior to fuel injection, the swirl forming means applies lateral swirl into the combustion chamber (in the cylinder) during the intake stroke in the cylinder axial pressure distribution (ignition on the spark plug side). With a high swirl strength and a pressure distribution with a small swirl strength on the piston side, so that an axial pressure difference occurs at the center of the combustion chamber (the swirl pressure of the larger swirl strength is smaller than that of the lower swirl strength). The fuel pressure is conveyed to the vicinity of the spark plug due to the pressure difference, and the fuel spray is concentrated on the spark plug side to realize a stratified combustion mode.

Therefore, even in a flat piston, stable combustion at a lean air-fuel ratio at a low load is realized, and the fuel consumption is greatly reduced.
R can be added. As a result, exhaust gas, especially N
Ox can be reduced.

[0017]

Embodiments of the present invention will be described with reference to the drawings. The same reference numerals in each drawing indicate the same or common elements.

FIG. 1 is an explanatory view according to a first embodiment of the present invention. FIG. 1 (a) shows a main part of the ignition type gasoline engine cylinder (only one is taken out) and a swirl. FIG. 2B is a plan view showing the arrangement relationship between the forming valve 8 and the secondary intake passage 11a, and FIG. 2B is a configuration diagram of the engine system according to the embodiment as viewed from the front.

The cylinder head 14 has two intake valves 1
6a and 16b (when collectively expressing 16a and 16b, the reference numeral 16) and exhaust valves 17a and 17b (17a and 17b)
17b is collectively represented by reference numeral 17), and the ignition plug 2 is arranged so as to face the combustion chamber 13 at the center thereof. The intake port of the intake pipe 11
The two intake valves 16a and 16b are divided into two by a partition wall 110 corresponding to the two intake valves.
1-1, and the other is indicated by reference numeral 11-2.

The partition wall 1 downstream of the throttle valve 10 in the intake pipe 11
A valve (hereinafter referred to as a diversion valve) 8 that is opened and closed according to a combustion mode (lean or stoichiometric) according to the load state is provided immediately upstream of the intake pipe 11. A secondary intake passage 11a that bypasses the flow dividing valve 8 is provided. The secondary intake passage 11a is directed to one of the two intake valves 16b, so that the outlet of the secondary intake passage 11a faces the intake port 11-1 corresponding to the intake valve 16b. Is set. The secondary intake passage 11a can be formed inside the intake pipe in addition to the independent passage provided outside the intake pipe as shown in the figure.

In this embodiment, the two intake valves 16 described above are used.
a, 16b, the diverting valve 8, and the secondary intake passage 11a allow a swirl forming means (described later) to transfer a horizontal swirl airflow swirling around the cylinder axis into the combustion chamber 13 during the intake stroke in the lean burn mode. A swirl forming means for forming a swirl intensity distribution in the cylinder axial direction such that the swirl intensity on the spark plug 2 side is greater than that on the piston 6 side).

FIG. 7A shows that the opening of the flow dividing valve 8 is changed from 0 to 1.
This graph shows the lateral swirl intensity characteristics of the intake air in the combustion chamber 13 generated when the value is changed to 00%.

When the opening of the flow dividing valve 8 is 0, it is in a fully closed state. In this case, the intake air is exclusively supplied to one of the intake valves 1 through the secondary passage 11a of the intake pipe 11 and the intake port 11-1.
6b, the air flows into the combustion chamber 13 so that a horizontal swirl occurs in which the air flows along the cylinder wall of the cylinder 7,
Further, the swirl strength is maximized. As the flow dividing valve 8 is opened, the degree of air flowing from the main passage 11 increases, and the swirl strength decreases. The swirl strength is determined by the distribution of the airflow passing through the intake valves 16a and 16b.

In the drawing, reference numeral 5 denotes an air amount detection sensor disposed upstream of the throttle valve 10, reference numeral 15 denotes a pressure sensor in the intake pipe,
10A is a throttle sensor. The opening of the throttle valve 10 is drive-controlled through a controller 12 and a throttle actuator (for example, a stepping motor) 9 based on a signal from an accelerator opening sensor (not shown). The controller 12 detects the load of the engine based on the detection signals of the throttle sensor 10A and the intake pipe pressure sensor 15, and controls the opening of the flow dividing valve 8 via the stepping motor 19 based on the detected load.

The intake air passes through an air flow detection sensor 5, a throttle valve 10, and an intake valve 16 in an intake pipe 11, and is taken into a combustion chamber 13 of the engine.

FIG. 7B shows a control pattern of the flow dividing valve 8 during a low load operation (lean burn mode) and an intensity control pattern of the lateral swirl generated in the combustion chamber 13 at that time.

During low-load operation, when the piston is at the intake top dead center (start point of the intake stroke), the opening of the shunt valve 8 is large, and the opening decreases with the progress of the intake stroke. Near the position (end point of the intake stroke), the flow dividing valve 8 is closed.

By controlling the flow dividing valve 8 as described above, a horizontal swirl airflow swirling around the cylinder axis is introduced into the combustion chamber 13 during the intake stroke in the lean burn mode in the latter half of the intake stroke of the engine. It is possible to construct a swirl forming means for forming the swirl so as to increase the strength. Further, by such a swirl formation, the air flow of the lateral swirl in the combustion chamber 13 is formed such that the swirl strength on the spark plug side is larger than that on the piston side. That is, the air with low swirl strength that has flowed into the combustion chamber first affects the flow near the piston,
The swirl-strength air that has flowed in later affects the flow near the cylinder head (near the spark plug). Therefore, by controlling the opening of the flow dividing valve 8 as the intake stroke progresses, a desired air flow distribution can be formed in the cylinder.

FIG. 6 shows the principle of stratification of the air-fuel mixture in the combustion chamber in this embodiment.

As described above, by increasing the intensity distribution of the horizontal swirl of the air in the cylinder (combustion chamber) 13 from the piston 6 toward the spark plug 2 (see FIG. 6A),
The higher the swirl strength, the more the air is pressed against the inner wall of the cylinder by centrifugal force. Therefore, the pressure near the ignition plug 2 at the center of the swirl decreases. FIG. 6B shows the strength characteristics of the lateral swirl in the cylinder axis direction at this time. Since a pressure difference distribution in which the pressure decreases from the piston 6 toward the spark plug 2 is generated, if fuel is injected by the injector 1 after such a swirl intensity distribution is formed, the air-fuel mixture (fuel spray) 3 ′ moves in the spark plug direction. It is conveyed toward. Since ignition occurs near the compression top dead center, it is necessary to increase the swirl strength toward the spark plug during the compression stroke. As shown in FIG. 6C, there is a correlation between the distribution of the swirl intensity during the intake stroke and the distribution of the swirl during the compression stroke. Therefore, the swirl intensity distribution during the compression stroke is controlled by controlling the swirl intensity during the intake stroke. Can be similarly formed.

FIG. 5 shows a state in which fuel injection is performed during the compression stroke after forming the above-described lateral swirl intensity distribution in the cylinder axis direction. FIG. 5A shows a state in which fuel 3 is injected by the injector 1 in the first half of the compression stroke.
(B) shows a state in which the fuel spray 3 ′ in the latter half of the compression stroke is concentrated near the ignition plug 2.

As shown in FIGS. 5 (a) and 5 (b), a horizontal swirl air flow with the above-mentioned pressure distribution is formed in the cylinder, the fuel is trapped by the swirl air, and the fuel is collected in the spark plug. In this case, it is necessary to reduce the spray speed as the fuel spray. If the spray speed is high, the injected fuel collides with the opposed wall surface. For example, if the injection speed is 2
Spray distance of 5m / s or less or 4ms after injection
0 mm or less. The fuel particle size is less than 20 microns, preferably less than 10 microns.

According to the present embodiment, a stratified charge combustion can be performed even with a flat piston to realize a super-lean burn, so that the operable air-fuel ratio can be widened, thereby improving fuel efficiency and exhaust gas. Purification can be improved.

Further, when homogeneous combustion is required as in the stoichiometric air-fuel ratio mode such as a medium or high load operation, the flow dividing valve 8 is fully opened. By this full opening control, the intake valve 16
In the combustion chamber 13, a tumble air flow is formed during the intake stroke in which a substantially equal amount of air flows through a and 16 b. By performing fuel injection during this intake stroke, a well-homogenized air-fuel mixture is formed, and sufficient homogeneous combustion can be realized.

FIG. 8 shows a configuration diagram of a direct injection engine according to a second embodiment of the present invention.

In this embodiment, basically, the swirl airflow swirling around the cylinder axis is introduced into the combustion chamber 13 during the intake stroke in the rich burn mode (low load operation).
By forming the swirl intensity in the latter half of the intake stroke of the engine to be larger than that in the first half of the intake stroke, a horizontal swirl having the same swirl intensity distribution as in the first embodiment is formed. Is configured as follows. The engine is an ignition gasoline engine using a flat piston 6. Cylinder head 1
4 has two intake valves 1 in a layout as shown in FIG.
6a and 16b, and not shown in FIG.
Two exhaust valves 17a and 17b are provided.

The intake valves 16a and 16b can be independently controlled for opening and closing operations by a set of electromagnetic solenoids 18 and 19, respectively, and the difference between the opening degrees of the intake valves 16a and 16b during the intake stroke in the lean burn mode. In the first half of the intake stroke (including the same opening), and in the latter half of the intake stroke, the opening of the other intake valve 16b is controlled to be larger than that of the one intake valve 16a (here, the intake valve 16a is closed, The intake valve 16b is opened) to constitute the swirl forming means.

The electromagnetic solenoids 18 and 19 are arranged in the axial direction. Among them, the solenoid 18 is for opening the intake valve 16, and the solenoid 19 is for closing the intake valve 16. Is the casing 2 together with the return springs 23 and 24 for opening and closing assistance.
8 interior. A gap is provided between the electromagnetic solenoids 18 and 19, and a movable magnetic body 22 integral with the shaft of the intake valve 16 is interposed in this gap.

When a voltage is applied to the solenoid 18 via the drive circuit 30, for example, the intake valve 16, ie, 16a, 16b, is magnetically attracted to the solenoid 18 via the movable magnetic body 22 and opens. When a voltage is applied to the solenoid 19 in the same manner, the solenoid 19
Magnetically attracted to the side and the valve closes.

The exhaust valves 17, that is, 17a and 17b have the same configuration. The solenoid 20 is for opening the exhaust valve 17, and the solenoid 21 is for closing the exhaust valve 17. These solenoids 20 and 21 are housed in a casing 29 together with return springs 26 and 27 for opening and closing assistance. . Electromagnetic solenoids 20, 21
A gap is secured between them, and a movable magnetic body 25 integral with the shaft of the exhaust valve 17 is interposed in this gap.

The fuel is supplied from the injector 1 which can directly inject fuel into the cylinder 7. The injector 1 is driven by the drive circuit 32. The throttle valve 10 is a motor 9
The opening and closing operation is performed by the throttle sensor 10.
A.

The accelerator opening α is detected by an accelerator opening sensor (not shown), and in this embodiment, the intake and exhaust valves are controlled by capturing the load at least based on the accelerator opening sensor signal. In addition, the controller 12 controls a throttle valve and the like based on signals from the various sensors. Reference numeral 42 in FIG. 8 is an in-cylinder pressure detection sensor.

FIG. 9 shows the control operation of the air flow (horizontal swirl formation and forward tumbling) according to this embodiment.

FIG. 9A shows a state in which only one of the two intake valves 16a and 16b is controlled to be opened at the time of a low load condition. In this case, combustion is performed only from one of the intake valves 16b. Air flows into the chamber 13 (cylinder 7), and a horizontal swirl airflow can be formed in the combustion chamber 13.

FIG. 9B shows a method of controlling the air flow when the load increases (for example, the stoichiometric air-fuel ratio). As the load increases, more air is required. Therefore, in addition to the intake valve 16b, the other intake valve 16a is opened. As a result, a flow opposite to the swirl flowing from 16b is introduced, so that the swirl is weakened. Eventually, both of the two intake valves are opened. Thereby, the air flow in the cylinder in the combustion chamber 13 is determined by the shape of the intake port. In general, because of the improved output of the engine and the ease of layout of the engine and the intake port, those having an intake port shape in which a forward tumble is formed are widely used in automobile engines. Opening the valves 16a, 16b creates a forward tumbling airflow. By order tumble,
Mixing of air and fuel in the cylinder is promoted, and the air utilization rate is improved. Further, in a region where the load is large, the fuel is injected during the intake stroke to prolong the mixing time of the air and the fuel.

FIG. 10 shows another intake port shape applied to this embodiment. An intake port 11-2 corresponding to one of the two intake valves 16a and 16b is
It is formed on a curved line (helical line) that facilitates swirl.

FIG. 11 shows the intake valve 16 of this embodiment at a low load.
The control method at the time of the intake stroke of a and 16b is shown. By closing the other intake valve 16a earlier than one intake valve 16b, a lateral swirl can be generated in the combustion chamber in the latter half of the intake stroke as shown in FIG. 9 (a). In addition, by adding inertia operation of the air flow to this control, a weak lateral swirl can be generated near the piston and strong near the spark plug.

FIG. 12 shows an example of an intake valve control method during the intake stroke in the lean burn mode when the intake ports of FIG. 10 are employed for the independently controllable intake valves 16a and 16b as in this embodiment.

As in FIG. 11, swirl can be generated in the latter half of the intake stroke by closing the other intake valve 16a earlier than the other intake valve 16b. The one intake port 11-2 has a shape (helical port) that generates swirl, so that the two intake valves 16a,
Even if 16b is open, a weak horizontal swirl can be generated.

According to the present embodiment, by generating a lateral swirl that is weaker than at the beginning of the intake stroke, one of the intake valves 16a
When a large horizontal swirl is suddenly generated when the valve is closed, disturbance can be prevented, and a horizontal swirl having the intended pressure difference distribution can be generated smoothly. This makes it possible to generate a weak swirl near the piston and a strong swirl near the spark plug.

FIG. 13 shows an example of another intake valve control method using the in-cylinder injection engine system of FIG.

In this example, at the beginning of the intake stroke, the intake valve 16b
And a slight delay time is provided to open the intake valve 16a to generate a weak lateral swirl in advance. Thereafter, the two intake valves 16a and 16b are opened, and only the intake valve 16a is quickly closed in the latter half. This makes it possible to form a weak swirl flow near the piston and a strong swirl flow near the spark plug without using a helical port.

FIG. 14 shows a third embodiment of the present invention.

The present embodiment differs from the embodiment shown in FIG. 8 in that a lift adjusting mechanism 50 is added to the intake valve 16 in this embodiment. By providing the lift adjusting mechanism 50, the lift of the intake valve 16 can be made variable.
The lift adjusting mechanism 50 can be configured by setting the thickness of a stopper member provided at the upper end of the valve opening electromagnetic solenoid 18. In this example, the lift adjustment mechanism 50 is connected to the intake valve 16.
By providing one of the intake valves a and 16b, the lift of the other intake valve is controlled to be different from that of the one intake valve. As a result, one of the airflows becomes stronger, and a horizontal swirl airflow can be formed in the cylinder.

FIG. 15 shows an example of a drive pattern of the intake valve and the exhaust valve at a low load according to the present embodiment. FIG. 16 shows a control of the intake valves 16a and 16b at a low load in the present embodiment during the intake stroke. And the swirl intensity generated thereby.

As shown in FIGS. 15 and 16, the intake valves 16a,
16b is opened and closed using an independent electromagnetic opening and closing mechanism. For example, the lift of the intake valve 16a is smaller than that of the intake valve 16b, and the intake valve 16a is connected to the intake valve 16b.
It closes faster than it does. In this way, by changing the valve lift and opening / closing timing of the intake valves 16a and 16b by the lift adjusting mechanism 50, a weaker horizontal swirl can be generated in the first half as the air flow pattern in the combustion chamber, and a stronger horizontal swirl can be obtained in the second half. The lateral swirl strength can be generated, and the stratified combustion in which the fuel spray is concentrated near the ignition plug by forming the lateral swirl pressure distribution similar to the above-described embodiment can be performed. Also, when the intake valves 16a and 16b are suddenly operated by the electromagnetic mechanism, there is an effect of improving the efficiency of charging the intake air. The exhaust valve can also be controlled by an electromagnetic mechanism. By changing the opening / closing timing of the exhaust valve, the amount of residual gas in the cylinder is controlled, and NOx can be reduced by the internal EGR rate.

FIG. 17 shows a modification of the third embodiment.

FIG. 17A is similar to FIGS. 15 and 16 for the intake valves 16a and 16b, and is similar to the exhaust valves 17a and 17b.
About b, what was mechanically operated by the usual lift using the conventionally well-known camshaft.

In FIG. 17B, the intake valves 16a, 16b
Is an example in which the valve lift is smoothly changed instead of the on-off operation. By smoothly changing the valve lift, a shock when the valve is seated on the valve seat is reduced, and there is an effect of reducing seating noise.

In this example, the electromagnetic configuration is described as the intake valve. However, the effects of the present invention can be obtained by a hydraulic type or a mechanical type in which the opening / closing timing of the valve and the lift are made variable.

FIG. 18 shows two intake valves 16 by a hydraulic mechanism.
That is, an embodiment (fourth embodiment) in which the driving of 16a and 16b is controlled independently is shown.

Each of the intake valves 16a and 16b is provided with a hydraulic cylinder 52 and a return spring 51 for opening and closing the intake valves. The hydraulic cylinder 52
Connected to a hydraulic pump 72 via a valve 55;
It is connected to an oil return system passage via a valve 56. By opening the valve 55, closing the valve 56, and driving the hydraulic pump 72, hydraulic pressure is supplied to the hydraulic cylinder 52,
The intake valve 16 is driven to open. When closing the intake valve 16, the hydraulic pressure is reduced by closing the valve 55 and opening the valve 56. When the hydraulic pressure is reduced, the intake valve is returned by the spring 51. The opening / closing timing of the intake valve 16 and the lift can be controlled by controlling the timing of applying and removing the hydraulic pressure.

FIG. 19 shows the intake valves 16a and 16b of this embodiment.
This is an example in which the lift and opening / closing timing are changed. The lift of one intake valve 16a is made smaller than that of the other intake valve 16b, and the intake valve 16a is closed earlier than 16b in the latter half. This makes it possible to form a weak swirl in the early stage of intake and a strong lateral swirl in the second half. That is, a flow with a weak horizontal swirl near the piston and a strong flow with the horizontal swirl near the spark plug can be formed.

FIG. 20 shows an example of the control of the air flow pattern in the combustion chamber (cylinder) of the engine which is enabled by the above embodiment.

When the engine torque and the engine speed are small, that is, at the time of a so-called low load, in the super lean burn operation in which the air-fuel ratio is 40 or more, it is necessary to concentrate the air-fuel mixture near the spark plug (stratification). This stratified combustion is made possible by forming a lateral swirl airflow with a swirl intensity difference distribution in the axial direction in the engine. In this case, EGR can be added as needed. In the operation in which EGR is added at the air-fuel ratio of 20 to 40 when the load is increased, if the air-fuel mixture is concentrated too much around the spark plug, oxygen shortage occurs and smoke is easily generated. Therefore,
The air flow in the cylinder is set as a lateral swirl so that the air-fuel mixture in the cylinder is not relatively concentrated. The transverse swirl flow is effective in promoting the mixing of air and fuel because the flow is easily preserved even when the piston approaches the compression top dead center. Further, even in the case where EGR is added, there is also an effect of improving the mixture of the EGR and the air-fuel mixture and stabilizing the combustion.

As the load increases, the stoichiometric air-fuel ratio (14.
In 7), a weak swirl flow is obtained even in the operation with EGR added (homogeneous stoichiometric weak swirl).

When the load is further increased, it is necessary to introduce a large amount of air in order to increase the output in an operation state in which EGR is not added, so that the resistance of the intake valve, the intake pipe, etc. is not increased, and the shape of the intake pipe is reduced. Flow determined by To improve the output, it is important to promote the mixing of air and fuel in the cylinder and increase the air utilization rate. The forward tamper (homogeneous stoichiometric tumbling) facilitates the mixing of air and fuel.

FIG. 21 is an operation block diagram of the air flow control according to the present invention.

A target engine torque (load) is calculated by a signal of an accelerator opening sensor, a vehicle speed, a gear position, an engine speed, an air amount sensor, and the like.

The air flow mode and the air-fuel ratio in the cylinder are selected according to the target engine torque and the engine speed.

The air flow modes in the cylinder are the horizontal swirl and forward tumbling described above, and are selected according to the operating conditions (engine torque, engine speed). After selecting the air flow mode in the cylinder, the air flow control parameter in the cylinder is calculated, and the air flow is controlled by the air flow control means in the cylinder (that is, the horizontal swirl and forward tumble forming means described in the previous embodiment). . For example, FIG.
In this embodiment, the air flow control means in the cylinder is the intake valves 16a, 16b or the secondary intake passage 11a provided in the intake port, and the flow dividing valve 8. The air flow control parameters in the cylinder are the opening / closing timing of the intake valve, Lift and diverting valve opening. The fuel injection amount is determined according to the engine torque, and the optimum fuel injection timing is calculated according to the air flow pattern.

FIG. 22 shows an example in which the flow dividing valve 8 of FIG. 1 is used as the air flow control means in the cylinder. The degree of opening and timing of the flow dividing valve are calculated according to the air flow mode in the cylinder, and the flow dividing valve actuator (for example, a stepping motor) is controlled.

FIG. 23 is a flowchart. A target engine torque is calculated from an accelerator opening, a vehicle speed, a gear position, an engine speed, an air flow meter, and the like. The air flow mode, air-fuel ratio, and target air amount in the cylinder are calculated according to the target engine torque and the engine speed. The air flow modes in the cylinder are swirl and forward tumble, and are selected according to operating conditions (engine torque, engine speed). The opening / closing timing of the flow dividing valve is calculated, and the flow dividing valve actuator is driven. If the position is not the target value, the opening / closing timing and opening are corrected. According to the engine torque, the fuel injection pulse width and the injection timing are calculated, the combustion fluctuation is detected by the in-cylinder pressure sensor and the rotation speed sensor, and if it exceeds the allowable value, it is said that stratification is insufficient. Correct the diversion valve opening so that the strength of the lateral swirl increases.

FIG. 24 shows an example in which the air flow control means (horizontal swirl forming and forward tumble forming means) is an intake valve (variable valve) which can be independently opened and closed as shown in FIGS. The in-cylinder air flow control parameters are the opening / closing timing of the intake valve and the lift.

FIG. 25 is a flowchart. A target engine torque is calculated from an accelerator opening, a vehicle speed, a gear position, an engine speed, an air flow meter, and the like. The air flow mode, air-fuel ratio, and target air amount in the cylinder are calculated according to the target engine torque and the engine speed. The air flow modes in the cylinder are swirl and forward tumble, and are selected according to operating conditions (engine torque, engine speed). The opening / closing timing and lift of the two intake valves are calculated, and the intake valves are driven. The air amount for each cylinder to the engine 13 is controlled.

The position of the opening of the intake valve is detected by a position sensor, and feedback control is performed to determine whether the opening and closing are controlled at the target valve position and timing. The amount of air sucked into the engine is detected by an air amount sensor for each cylinder, and compared with a target air amount to perform feedback control. Further, the output torque of the engine is detected by a crank angle sensor or an in-cylinder pressure sensor, and it is compared with the target engine torque to perform feedback control. When the in-cylinder pressure sensor is used, the amount of air in the cylinder can be detected from the in-cylinder pressure after the intake valve is closed.
The air flow meter can be eliminated. The air flow mode can be arbitrarily changed by operating the intake valve on one side in the cylinder air flow mode.

[0077]

According to the present invention, super-lean burn is realized by enabling stratified combustion even with a flat piston, and an in-cylinder injection type engine capable of widening the operable air-fuel ratio is realized, thereby improving fuel efficiency. Thus, the exhaust can be reduced.
Further, since a homogeneous mixture can be formed in the medium and high load operation regions, the engine output can be improved.

[Brief description of the drawings]

FIG. 1 is an explanatory view according to a first embodiment of the present invention, in which (a) is a main part of a cylinder (only one is taken out) of an ignition type gasoline engine and an on-off valve for forming a swirl; 8 is a plan view showing an arrangement relationship between the engine system and the secondary intake passage 11a, and FIG. 8B is a configuration diagram of the engine system according to the present embodiment as viewed from the front.

FIG. 2 is an explanatory diagram showing a conventional air-fuel mixture formation state of in-cylinder injection.

FIG. 3 is an explanatory view showing a state in which a mixture is formed by in-cylinder injection using a flat piston.

FIG. 4 is an explanatory diagram showing engine torque characteristics of a flat piston and a cavity-equipped piston.

FIG. 5 is an explanatory diagram showing a state in which a horizontal swirl airflow is formed and fuel is injected according to the embodiment.

FIG. 6 is an operation explanatory view of the embodiment.

7A is a characteristic diagram showing a relationship between the opening degree of the flow dividing valve and the lateral swirl used in the above embodiment, and FIG. 7B is a graph showing the control of the flow dividing valve in the intake stroke of the above embodiment during a low load operation and the state at that time; Explanatory drawing which shows the state of horizontal swirl formation in a combustion chamber.

FIG. 8 is a system configuration diagram according to a second embodiment of the present invention.

FIG. 9 is an explanatory diagram showing a relationship between control of two intake valves and air flow in the second embodiment.

FIG. 10 is an explanatory diagram showing an alternative regarding the intake port shape of the second embodiment, and showing the relationship between control of two intake valves and air flow when this alternative is used.

FIG. 11 is an explanatory diagram showing a relationship between intake valve control and in-cylinder swirl strength of the second embodiment.

FIG. 12 is an explanatory diagram showing a modified example showing a relationship between intake valve control and in-cylinder swirl strength of the second embodiment.

FIG. 13 is an explanatory diagram showing another modified example showing the relationship between the intake valve control and the swirl strength in the cylinder according to the second embodiment.

FIG. 14 is a system configuration diagram according to a third embodiment of the present invention.

FIG. 15 is an explanatory diagram showing an operation example of an intake valve and an exhaust valve of the third embodiment.

FIG. 16 is an explanatory diagram showing a relationship between intake valve control and in-cylinder swirl strength of the third embodiment.

FIG. 17 is an explanatory view showing a modification of the operation of the intake valve and the exhaust valve of the third embodiment.

FIG. 18 is a system configuration diagram according to a fourth embodiment of the present invention.

FIG. 19 is an explanatory diagram showing a relationship between intake valve control and in-cylinder swirl strength of the fourth embodiment.

FIG. 20 shows an example of control of an air flow pattern in a combustion chamber (cylinder) of an engine which is enabled by the above embodiment.

FIG. 21 is a control block diagram for executing the above embodiments.

22 is a block diagram in the case where the flow dividing valve of FIG. 1 is used to form air flow in the cylinder of FIG. 21.

FIG. 23 is a flowchart of engine control when the above-mentioned flow dividing valve is used.

FIG. 24 shows the formation of air flow in the cylinder of FIG.
FIG. 14 is a block diagram in the case of using 14 independent electromagnetic control type intake valves (variable valves).

FIG. 25 is a flowchart of engine control when the variable valve is used.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 ... Injector, 2 ... Spark plug, 3 ... Fuel spray, 6
... Piston, 7 ... Cylinder, 8 ... Diversion valve (swirl forming means), 10 ... Throttle valve, 11 ... Intake passage, 11a ... Secondary intake passage (Swirl formation means), 12 ... Controller,
13: combustion chamber, 16 (16a, 16b): intake valve, 17
(17a, 17b) ... exhaust valves, 18, 19 ... electromagnetic solenoids (swirl forming means, air flow forming means).

Continuation of the front page (51) Int.Cl. ⁷ Identification symbol FI Theme coat II (reference) F02D 41/02 325 F02D 41/02 325F 325G (72) Inventor Yoko Nakayama 7-1-1, Omikamachi, Hitachi City, Ibaraki Prefecture Inside Hitachi Research Laboratory, Hitachi, Ltd. (72) Noboru Tokuyasu 7-1-1, Omikamachi, Hitachi City, Ibaraki Prefecture Inside Hitachi Research Laboratory, Hitachi, Ltd. (72) Inventor Minoru Osuka 7, Omikamachi, Hitachi City, Ibaraki Prefecture No. 1-1 Inside Hitachi, Ltd. Hitachi Research Laboratory (72) Inventor Yoshihiro Sukekawa 7-1-1, Omika-cho, Hitachi City, Ibaraki Prefecture F-term in Hitachi, Ltd. Hitachi Research Laboratory F-term (reference) 3G023 AA02 AA05 AB03 AC04 AD01 AD05 AD07 AG01 AG02 AG03 3G092 AA01 AA06 AA09 AA10 AA11 AA13 AA16 AA17 BA01 BA05 BA06 BA07 BA08 BB01 DA01 DA07 DA12 DA14 DC03 DC06 DC07 DE03S DF01 DF04 DF05 DG05 DG09 EA01 EA03 EA07 GA01 HA11Z HA13X HA1 3Z HE01Z HE03Z HE05Z HE06Z 3G301 HA01 HA06 HA13 HA16 HA17 JA02 JA24 JA25 KA08 KA09 LA03 LA05 LA07 LA08 LB04 LC01 LC03 LC08 MA01 MA11 MA18 NB11 ND01 NE01 NE11 PA01A PA01Z PA07Z PA11Z PA17Z PC01Z PE01ZPEZZPEZZ PEZZ PEZ PEZZ PEZZ

Claims (6)

[Claims] An injector for directly injecting fuel into a combustion chamber formed between a cylinder head and a piston;
An in-cylinder injection type engine having a spark plug disposed in a cylinder head, wherein a lateral swirl airflow swirling around a cylinder axis is introduced into a combustion chamber during an intake stroke in a lean mixture combustion mode,
An in-cylinder injection engine comprising a swirl forming means for forming a swirl intensity distribution in a cylinder axial direction such that a swirl intensity on a spark plug side is higher than that on a piston side. 2. An injector for directly injecting fuel into a combustion chamber formed between a cylinder head and a piston,
An in-cylinder injection type engine having a spark plug disposed in a cylinder head, wherein a lateral swirl airflow swirling around a cylinder axis is introduced into a combustion chamber during an intake stroke in a lean mixture combustion mode,
An in-cylinder injection engine comprising a swirl forming means for forming a swirl intensity in a latter half of an intake stroke of the engine to be larger than that in a first half of the intake stroke. 3. The cylinder head is provided with two intake valves, and a downstream side of a throttle valve in the intake passage is provided with a valve that is opened and closed in accordance with a load state of an engine, and a secondary valve that bypasses the valve. The in-cylinder injection engine according to claim 1 or 2, wherein the swirl forming means is configured by arranging an intake passage so as to face one of the two intake valves. 4. The cylinder head is provided with two intake valves, and these intake valves are capable of independently controlling an opening / closing operation. The two intake valves during an intake stroke in a lean air-fuel mixture combustion mode are provided. The swirl forming means is configured such that a valve opening difference is reduced in the first half of the light absorption stroke in the first half of the light absorption stroke, and the opening of the other intake valve is controlled to be larger in the second half of the intake stroke as compared with the one of the intake valves. The in-cylinder injection engine according to claim 2. 5. The cylinder head is provided with two intake valves, and these intake valves are capable of independently controlling an opening and closing operation, and the two intake valves are used during an intake stroke in a lean air-fuel mixture combustion mode. 3. The direct injection engine according to claim 1, wherein the swirl forming means is configured by setting a timing of closing one valve earlier than that of the other intake valve. 6. The cylinder according to claim 3, wherein an intake port corresponding to one of the two intake valves is formed on a curved line that facilitates forming a horizontal swirl. Internal injection engine.