JPH10252512A - Compressed ignition internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To spread a stable combustion region in a compressed ignition internal combustion engine. SOLUTION: A maximum injection amount without generating knocking when fuel is injected only from a port injection valve 15 is set as a reference injection amount, its corresponding load amount is set to a reference load amount, in a load region of the reference load amount or less, fuel of total amount of necessary injection is injected from the port injection valve 15 during an intake stroke on and after the top dead center of a piston 5, and so as to increase a valve open overlap period of an intake/exhaust valve 8, 9 in accordance with decreasing of a load, an open/close timing variable means is controlled. An internal EGR amount is increased in accordance with decreasing of a load, an intake heating effect is increased, self ignitability in a light load region is improved, a stable combustion region in a light load side is spread from in the past.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a compression ignition type internal combustion engine which compresses a premixed gas at high compression and ignites it.

[0002]

2. Description of the Related Art A premixed compression ignition gasoline engine, which is one type of compression ignition internal combustion engine, has attracted attention as a candidate for an engine that simultaneously achieves high efficiency and cleanliness. In a premixed compression ignition gasoline engine, as in the conventional spark ignition gasoline engine, gasoline is injected into the intake port to introduce a completely premixed gas into the cylinder and compress it to a high compression ratio comparable to a diesel engine, thereby igniting. The compression ignition is performed without using a plug.

According to the literature ("Academic Lectures Preprints 951 1995-5" of the Society of Automotive Engineers of Japan), in a premixed compression ignition gasoline engine, the air-fuel ratio is within a predetermined range (for example, 34 to 4).
According to 4), stable compression ignition combustion of a premixed gas is obtained, and low fuel consumption comparable to that of a diesel engine and a low NOx level of less than one tenth of a diesel engine can be achieved. Further, it is said that the generation of NOx in the cylinder can greatly reduce the NOx concentration in the case of the homogeneous mixture as compared with the stratified intake.

[0004] However, when the air-fuel ratio is leaner than the predetermined range, ignition failure is more likely to occur, and when the air-fuel ratio is richer, knocking is more likely to occur. Therefore, the operating range of stable combustion is limited to a narrow range. there were.

[0005] The cause of the ignition failure on the lean side is that the self-ignition temperature of gasoline is higher than that of light oil. Therefore, it was considered to increase the intake air temperature by increasing the intake air temperature to extend the lean limit of the stable combustion region.However, although the increase of the lean limit was achieved, the rich limit of the stable combustion region was shifted to the lean side to increase the lean limit. Knocking was apt to occur in the load region, and the stable combustion region was not expanded.
In addition, there has been a problem that the output is reduced due to a reduction in volumetric efficiency due to intake air heating.

Therefore, Japanese Patent Application Laid-Open No. 7-332140 discloses an intake port injection system which is provided with a port injection valve for injecting fuel into an intake port and an in-cylinder injection valve for directly injecting fuel into a cylinder. By injecting fuel in combination with the direct injection in the cylinder, the stable combustion range was expanded.

[0007]

In the technique disclosed in the above publication, the operation timing of the in-cylinder injection valve is set to a crank angle before compression top dead center of 8 to 30 degrees (hereinafter, the crank angle is abbreviated as ゜ CA). ), And the fuel injection amount (hereinafter referred to as the in-cylinder injection amount) Qd directly injected from the in-cylinder injection valve into the cylinder.
Is 15 to 25% of the total injection amount at full load. Here, the total injection amount Q is a sum of a fuel injection amount (hereinafter, referred to as a port injection amount) Qp injected from a port injection valve to an intake port and an in-cylinder injection amount Qd.

[0008] By the way, the total injection amount at full load is 15 to 25.
%, The in-cylinder injection amount Qd corresponds to the injection amount when the engine is not loaded. Therefore, on the light load side, the port injection amount Qp is small, and most of the fuel is in-cylinder injected from the in-cylinder injection valve. As a result, the actual combustion becomes similar to the stratified combustion by the direct injection in the cylinder, so that the exhaust gas cleaning which is a great advantage of the homogeneous charge compression ignition gasoline engine, that is, the low NOx level (one-tenth of the diesel engine) is achieved. The following is difficult to realize.

Further, in the technology disclosed in the above publication, the port injection amount Qp of the fuel at full load is as large as 75 to 85% of the total injection amount.

The present invention has been made in view of the above-mentioned problems of the prior art, and the problem to be solved by the present invention is to re-inhale the exhaust gas in the cylinder as the load becomes smaller on the light load side. The purpose of the present invention is to increase the intake air heating so as to prevent ignition failure from occurring, thereby expanding the stable combustion region on the light load side.

Another object of the present invention is to reduce the amount of in-cylinder injected fuel on the high load side to prevent knocking, thereby expanding the stable combustion area on the high load side. It is in.

[0012]

The present invention has the following features to attain the object mentioned above. (1) The present invention provides a cylinder, a piston reciprocating in the cylinder, a combustion chamber defined by the cylinder and the piston, and an intake valve for opening and closing an intake port for supplying intake air to the cylinder. An exhaust valve for opening and closing an exhaust port for exhausting gas in the cylinder; and an opening / closing timing variable means for changing the opening / closing timing of only the intake valve or both the intake valve and the exhaust valve of the intake valve and the exhaust valve. And a port injection valve that injects fuel into the intake port, and an in-cylinder injection valve that injects fuel directly into the combustion chamber, in a compression ignition type internal combustion engine, in which fuel was injected only from the port injection valve. When the maximum injection amount that does not cause knocking is set as the reference injection amount, the corresponding load amount is set as the reference load amount, and in a load region equal to or less than the reference load amount, the required injection amount is reduced. The opening / closing timing variable means controls such that the entire amount of fuel is injected from the port injection valve during the intake stroke after the top dead center of the piston, and the valve overlap period of the intake valve and the exhaust valve increases as the load decreases. This is a compression ignition type internal combustion engine.

Since the amount of fuel injected from the port injection valve is equal to or less than the amount of fuel at which knocking occurs, knocking does not occur even if the port injection fuel ignites itself. The smaller the load, the longer the valve opening overlap period of the intake valve and the exhaust valve. Therefore, the smaller the load, the more re-suction of exhaust gas in the cylinder (so-called internal EGR), and the higher the intake air is heated, Due to the rise in the gas temperature in the cylinder before the start of compression, the fuel reliably ignites even at a light load.

The fuel injection from the port injection valve is executed after the top dead center of the piston. This means that the fuel injection is executed after entering the substantial intake stroke. Can be prevented from flowing back through the intake port by the internal EGR gas, and the fuel is reliably sucked into the cylinder and mixed with fresh air to become a uniform premixed gas.

The opening / closing timing varying means may vary the opening / closing timing of only the intake valve, or may vary the opening / closing timing of both the intake valve and the exhaust valve. It suffices if the valve-opening period of the valve and the exhaust valve can be varied.

(2) The present invention also provides a cylinder, a piston reciprocating in the cylinder, a combustion chamber defined by the cylinder and the piston, and an intake port for supplying intake air to the cylinder. An intake valve that opens and closes, an exhaust valve that opens and closes an exhaust port for exhausting gas in the cylinder, a port injection valve that injects fuel into the intake port, and an in-cylinder that injects fuel directly into the combustion chamber In a compression ignition type internal combustion engine equipped with an injection valve, a maximum injection amount at which knocking does not occur when fuel is injected only from the port injection valve is defined as a reference injection amount, and a load amount corresponding thereto is defined as a reference load amount. In a load range exceeding the reference load amount, fuel corresponding to the reference injection amount out of the required injection amount is injected from the port injection valve, and A compression ignition type internal combustion engine is characterized in that an insufficient amount of fuel is injected from a direct injection valve before compression top dead center.

Since the amount of fuel injected from the port injection valve is equal to or less than the amount of fuel at which knocking occurs, knocking does not occur even if the port injected fuel self-ignites. The in-cylinder injected fuel injected from the in-cylinder injection valve evaporates near the compression top dead center,
The premixed gas containing the port injection fuel is mixed with the premixed gas while being cooled to form a combustible air-fuel mixture. In the fuel combustion process, a premixed gas including port injected fuel first self-ignites, and the in-cylinder injected fuel is ignited by the combustion gas to diffuse and burn. When the in-cylinder injected fuel performs diffusion combustion, the piston is descending, so the pressure rise in the cylinder is small,
Since the isothermal combustion continues, knocking does not occur at this time.

In the above inventions (1) and (2),
A cavity may be provided at the top of the piston, and the inside of the cavity may be used as a combustion chamber, or the top of the piston may be flat and a recess provided at the top of the cylinder head may be used as a combustion chamber.

In the above inventions (1) and (2), the fuel is not limited to gasoline, but may be a low cetane number light fuel to which an exhaust after-treatment device can be easily applied. is there.

[0020]

DESCRIPTION OF THE PREFERRED EMBODIMENTS One embodiment of a compression ignition type internal combustion engine according to the present invention will be described below with reference to FIGS.

FIG. 1 is a sectional view of a compression ignition type 4-cycle gasoline engine as a compression ignition type internal combustion engine. The engine 1 includes a cylinder block 2 and a cylinder head 3
And a piston 5 which can slide back and forth in the axial direction within the cylinder 4 and cannot rotate in the circumferential direction. The piston 5 rotates a crankshaft (not shown) via a connecting rod (not shown).

The cylinder head 3 is provided with an intake port 6 for supplying intake air into the cylinder 4 and an exhaust port 7 for exhausting gas in the cylinder 4. The intake port 6 is configured as a so-called swirl port that can generate a swirl. Note that a swirl control valve may be provided instead of the swirl port to generate swirl. The intake system is not provided with a throttle valve that is equipped with a spark ignition gasoline engine.

The intake port 6 opens at the inner top surface of the cylinder head 3 to form an intake port, and the intake port is controlled to be opened and closed by an intake valve 8. Exhaust port 7
Is opened at the inner top surface of the cylinder head 3 to form an exhaust port, and the exhaust port is controlled to be opened and closed by an exhaust valve 9. The intake valve 8 and the exhaust valve 9 are urged in a valve closing direction (a direction away from the piston 5) by a valve spring 10 provided between the cylinder head 3 and the piston 5 against the valve spring. The valve is opened by being moved in the approaching direction, and is opened and closed by a valve drive mechanism including the cam 11 and the like. In FIG. 1, a valve spring and a cam mechanism for the exhaust valve 9 are not shown.

A cavity 13 is formed at the top of the piston 5 and forms a combustion chamber 14 when the piston 5 is located at the top dead center. Cavity 1
Numeral 3 is arranged eccentrically from the center of the piston 5 so as not to interfere even when the intake valve 8 opens when the piston 5 is located at the top dead center.

The cylinder head 3 has a port injection valve 15 for injecting fuel into the intake port 6 with the injection hole facing the intake port 6 and a cylinder with the injection hole facing the combustion chamber 14. An in-cylinder injection valve 16 for directly injecting fuel into the combustion chamber 14 is provided.

Since the port injection valve 15 injects fuel into the low-pressure intake port 6, a low-pressure fuel injection valve can be used. For example, an injection valve that injects at a low pressure in the range of 0.15 to 1 MPa is used. Good. On the other hand, since the in-cylinder injection valve 16 is for injecting fuel into the high-pressure fuel chamber 14 when the piston 5 is located near the top dead center, it is necessary to use a high-pressure fuel injection valve. If the compression ratio is 14 to 20, an injection valve that injects at a high pressure in the range of 10 to 30 MPa is employed.

The fuel injection timing of the port injection valve 15 is set after the top dead center of the piston 5 during the intake stroke, and the fuel injection timing of the in-cylinder injection valve 16 is set before the compression top dead center of the piston 5. Have been.

The opening and closing timing of the intake valve 8 can be changed according to the operating conditions of the engine 1. Hereinafter, an example of the opening / closing timing varying means 12 for varying the opening / closing timing of the intake valve 8 will be described with reference to FIG.

The cam 11 for driving the intake valve 8 is fixed to an intake camshaft 20 rotatably supported by the cylinder head 3. The intake camshaft 20 is connected via a timing belt 23 to a pulley 21 that is driven to rotate by a crankshaft (not shown). The pulley 21 and the intake camshaft 20 are connected so as to rotate synchronously, and are connected so that their phases can be changed with each other.

More specifically, the end of the intake cam 20 is rotatably mounted inside the pulley 21.
A cylindrical plunger 22 is interposed between the pulley 21 and the outside of the end so as to be movable in the axial direction.

The outer peripheral surface of the plunger 22 and the pulley 2
1 has a helical spline 22a (pulley 21) that meshes with each other and is relatively slidable in the axial direction.
Of the helical spline is omitted). Due to the meshing of these helical splines, the rotational driving force of the pulley 21 is transmitted to the plunger 22.

Further, a helical spline 20a that meshes with each other and is relatively slidable in the axial direction is also provided on each of the outer peripheral surface of the intake camshaft 20 and the inner peripheral surface of the plunger 22.
(The symbols for the helical splines of the plunger 22 are omitted). The rotational driving force of the plunger 22 is transmitted to the intake camshaft 20 by the meshing of these helical splines.

That is, the rotational driving force of the pulley 21 is transmitted to the intake camshaft 20 via the plunger 22 by the engagement of the helical spline of the pulley 21 and the plunger 22 and the engagement of the helical spline of the intake camshaft 20 with the plunger 22. You.

Further, pressure chambers (not shown) are provided on both sides of the plunger 22 in the axial direction, and a hydraulic circuit (not shown) for supplying and shutting off the hydraulic pressure to each pressure chamber is provided. Is provided.

By controlling the hydraulic pressure applied to each pressure chamber, the plunger 22 can be moved in the axial direction within a predetermined range. When the plunger 22 moves in the axial direction, it moves while rotating along a helical spline that meshes with the pulley 21 and also moves while rotating along a helical spline that meshes with the intake camshaft 20.

Therefore, the plunger 22 is connected to the pulley 21
Is moved in the axial direction with respect to the pulley 21.
And the phase of the intake camshaft 20 can be continuously and continuously changed within a predetermined range. Furthermore, after the plunger 22 is moved to a predetermined position, the hydraulic pressure in the pressure chambers is maintained by shutting off the hydraulic circuit, and the plunger 22 is immobilized.
And the intake camshaft 20 can be maintained in a predetermined phase.

That is, the opening / closing timing of the intake valve 8 can be changed as desired by the opening / closing timing control means 12. The opening / closing timing of the exhaust valve 9 can be changed as desired by opening / closing timing control means of the same type as the opening / closing timing control means 12 of the intake valve 8.

Next, the operation of the engine 1 will be described. First, the engine 1 is actually operated in advance, and a maximum fuel injection amount that does not cause knocking when fuel is injected only from the port injection valve 15 is determined. This injection amount is set as a reference injection amount Qb. The load to be applied is set as a reference load Hb.

In the following description, as shown in FIG. 3, a load region below the reference load amount Hb is defined as a light load region, and a load region exceeding the reference load amount Hb is defined as a high load region. In the engine 1, in the light load region, the entire amount of fuel of the injection amount Qq required according to the load is injected from the port injection valve 15, and in the high load region, the injection amount Qq required according to the load is provided. Of the reference injection amount Qb is injected from the port injection valve 15, and the remaining amount Qd (Qd = Qq-Q
Fuel injection control is performed so that the fuel of b) is injected from the in-cylinder injection valve 16. Note that the load amount (the magnitude of the load) is determined from the accelerator opening and the engine speed.

Further, in the engine 1, the opening / closing timing control means 12 always sets the valve opening timing of the intake valve 8 to the most retarded angle (indicated by a solid line in FIG. 4) in the reference load amount Hb and the high load region. In the low load range, the valve opening timing of the intake valve is controlled so as to gradually and continuously shift to the advance side as the load decreases. That is, in the light load region, as shown in FIG. 5, the valve opening overlap period of the intake valve 8 and the exhaust valve 9 is controlled to be longer as the load is reduced. The compression ratio also increases.

Hereinafter, the combustion process in the engine 1 will be described separately for a light load region and a high load region. (1) Light Load Region In the intake stroke, there is a valve-open overlap period of the intake valve 8 and the exhaust valve 9 before and after the top dead center of the piston 5, and the cylinder 4 is closed during the valve-open overlap period before the piston 5 reaches the top dead center. The high-temperature exhaust gas remaining inside the air flows back into the intake port 6 and the piston 5
When the substantial intake into the cylinder 4 starts after the top dead center, the exhaust gas flowing back into the intake port 6 is sucked into the cylinder 4 again (hereinafter, this exhaust cylinder 4 is referred to as internal EGR). Naturally, the suction of the internal EGR into the cylinder 4 is performed before the fresh air is sucked.

In the light load range, the valve opening overlap period is longer as the load is smaller.
The amount is large. Then, since the heat amount of the internal EGR is used for intake air heating, the smaller the load, the higher the intake air temperature, and the higher the gas temperature in the combustion chamber 14 before the start of compression.
Furthermore, the substantial compression ratio is increased by the early closing of the intake valve 8, so that the gas temperature in the cylinder 4 at the compression end point increases.

In the light load region, the entire injection amount Qb required according to the load at that time is injected from the port injection valve 15, and no fuel is injected from the in-cylinder injection valve 9. The fuel injection timing of the port injection valve 15 is set after the top dead center of the piston 5 in the intake stroke, and this is nothing other than after the actual intake into the cylinder 4 is started. The fuel injected from the valve 15 is reliably sucked into the cylinder 4 without flowing backward in the intake port 8.

The internal EGR is first sucked into the cylinder 4 and then fresh air is sucked. At this time, since the intake port 6 is constituted by a swirl port,
Both the internal EGR and the fresh air are swirled to be sucked in a stratified state, and the internal EGR and the fresh air are not completely mixed in the cylinder 4.

By adjusting the fuel injection timing of the port injection valve 15 at the time of intake of fresh air after the internal EGR has been sucked into the cylinder 4, the fuel injected from the port injection valve 15 is uniformly predicted together with fresh air. Mixed, cylinder 4
Inside, it exists as a layer of uniform premixed gas above the internal EGR.

That is, even after the fuel sucked together with the fresh air is sucked into the cylinder 4, a large amount of oxygen still exists around the fuel. In addition to the fact that the fresh air is heated by the internal EGR, the cylinder 4
The interior becomes extremely self-igniting.

Therefore, when the combustion chamber 14 is compressed to a predetermined level in the subsequent compression stroke, the fuel is reliably self-ignited. Thereby, in the engine 1, the stable combustion region on the light load side is expanded as compared with the conventional compression ignition gasoline engine.

In the above description, it is assumed that most of the fuel is premixed with fresh air, and that the atmosphere around the fuel is oxygen-sufficient.
Stratification of internal EGR and fresh air is realized by swirl.

However, when attention is paid to the improvement of ignition performance caused by the rise in intake air temperature, the internal E
There is also a view that it is better to completely mix the internal EGR and fresh air in the cylinder 4 than to stratify the GR and fresh air. Therefore, it is also possible to generate a tumble in the intake air to completely mix the internal EGR, fresh air, and fuel, thereby heating the fresh air. When tumble is generated in this way,
Compared to the case where swirl is generated, although the oxygen concentration around the fuel becomes lower, the temperature of the intake air around the fuel rises, so that the self-ignition property of the fuel is improved as compared with the conventional compression ignition gasoline engine.

In the light load range, as the load increases, the internal EGR decreases and the degree of heating of the intake air also decreases, so that knocking does not occur even when approaching the reference load amount Hb.

(2) High load range In the high load range, although the intake valve 8 and the exhaust valve 9 have an overlapped valve opening period during the intake stroke, the valve opening timing of the intake valve 8 is set to the most retarded angle and the valve is opened. Since the valve overlap period is very short, there is very little internal EGR and, therefore, very little intake air heating.

In the high load region, fuel of the reference injection amount Qb of the injection amount Qq required according to the load is injected from the port injection valve 15, and the remaining amount Qd (Qd = Qq-
The fuel of Qb) is injected from the in-cylinder injection valve 16. Regarding the fuel injection timing, as described above, the port injection valve 15
Are set after the top dead center of the piston 5 in the intake stroke, and set before the compression top dead center from the in-cylinder injection valve 16.

Therefore, first, fuel is injected from the port injection valve 15 after the top dead center of the piston 5 in the intake stroke, and becomes fresh air and a uniform premixed gas in the cylinder 4. On the other hand, the in-cylinder injection valve 16 outputs the piston 15 before the compression top dead center.
The fuel is injected toward the cavity 13 of the fuel injection valve. This fuel injection timing is before the fuel injected from the port injection valve 15 self-ignites. Therefore, the fuel injected from the in-cylinder injection valve 16 is compressed. It evaporates near the dead center, and the pre-mixed air already sucked into the cylinder 4 is cooled by the vaporization latent heat. At the same time, the in-cylinder-injected fuel is mixed with the premixed air to obtain a combustible air-fuel ratio.

When the inside of the combustion chamber 14 is compressed to a predetermined degree near the compression top dead center in the compression stroke, the premixed gas self-ignites in the combustion chamber 14 and the fuel injected into the cylinder by the combustion gas releases the fuel. It ignites after top dead center and diffuses and burns.
Since the injection amount from the port injection valve 15 is the reference injection amount Qb, knocking does not occur when the port injected fuel self-ignites. Further, when the fuel injected into the cylinder performs diffusion combustion, the piston 5 is already descending, so that the pressure rise in the cylinder 4 is small and isothermal combustion is continued, so that knocking is avoided.

Therefore, in this engine 1, knocking can be avoided even in a high load region, and a stable combustion region on the high load side can be expanded as compared with the conventional compression ignition gasoline engine.

As described above, in this compression ignition type gasoline engine 1, the stable combustion region on the low load side is expanded, and in the low load region, uniform premixed combustion is performed to reduce the exhaust gas N.
Oxification can be realized. Therefore, a catalytic converter having a low NOx purification rate can be used in the exhaust gas purification system, and this is particularly effective when the present invention is applied to a large-displacement engine in which a light load to a medium load is a practical load range. .

Further, the light load region, which is the majority of the practical use region of the engine, is covered by port injection, and the in-cylinder injected fuel amount is small even in the entire engine use region.
Since the injection timing is near the compression top dead center, the dilution of the engine oil by the fuel hardly occurs.

In practicing the present invention, the compression ratio, the fuel injection timing of the port injection valve 15, the fuel injection timing of the in-cylinder injection valve 16, and the opening / closing timing of the intake valve 8 and the exhaust valve 9 are determined. The optimum value is experimentally determined and applied.

<Other Embodiments> In the above embodiment,
Although the opening / closing period of the intake valve 8 and the exhaust valve 9 is changed by changing the opening / closing timing of only the intake valve 8, FIG.
As shown in the above, the opening / closing period may be changed by changing the opening / closing timing of both the intake valve 8 and the exhaust valve 9. The means for changing the opening / closing timing of the exhaust valve 9 can be configured in the same manner as the aforementioned means for changing the opening / closing timing of the intake valve 8.

When the opening / closing timing variable means of the exhaust valve 9 is provided, during the warm-up operation when the engine water temperature is low, as shown in FIG. 7, the opening timing of the exhaust valve is gradually advanced as the engine water temperature decreases. By doing so, it is possible to enhance the intake heating effect as the engine water temperature is lower, and to improve the self-ignitability of the fuel injected through the port. This is because, if the valve opening timing of the exhaust valve 9 is advanced, the combustion gas in the cylinder 4 is exhausted before it expands sufficiently, and the expansion ratio becomes smaller.
R can be obtained.

[0061]

As described above, according to the present invention,
The maximum injection amount at which knocking does not occur when fuel is injected only from the port injection valve is set as the reference injection amount, and the corresponding load amount is set as the reference load amount. Injects the entire amount of fuel from the port injection valve during the intake stroke after the top dead center of the piston, and opens and closes so that the valve opening overlap period of the intake valve and exhaust valve increases as the load decreases By controlling the timing variable means, the smaller the load, the greater the internal EGR amount, the greater the intake air heating effect, the better the self-ignition properties in the light load range, and the more stable combustion range on the light load side than before. can do.

Further, in the low load region, uniform premixed combustion is performed by port injection, so that NOx reduction of exhaust gas can be realized.
A catalytic converter with a low purification rate can be used.

A maximum injection amount at which knocking does not occur when fuel is injected only from the port injection valve is set as a reference injection amount, and a load amount corresponding thereto is set as a reference load amount.
In the load range exceeding the reference load amount, the required injection amount of fuel corresponding to the reference injection amount is injected from the port injection valve. By injecting from the in-cylinder injection valve before the compression top dead center, occurrence of knocking in a high load region can be avoided, and the stable combustion region on the high load side can be expanded as compared with the related art. Further, since the amount of in-cylinder injected fuel can be reduced, the dilution of lubricating oil with fuel can be reduced.

[Brief description of the drawings]

FIG. 1 is a sectional view of a compression ignition type internal combustion engine according to an embodiment of the present invention.

FIG. 2 is a perspective view showing an example of an opening / closing timing variable unit used in the compression ignition type internal combustion engine of the present invention.

FIG. 3 is an example of a fuel injection map in the compression ignition type internal combustion engine of the present invention.

FIG. 4 is a diagram showing an example of opening and closing timing of intake and exhaust valves in a compression ignition type internal combustion engine of the present invention.

FIG. 5 is a diagram showing an example of a change in a valve-opening period of an intake / exhaust valve in a compression ignition type internal combustion engine of the present invention.

FIG. 6 is a diagram showing another example of the opening / closing timing of the intake / exhaust valve in the compression ignition type internal combustion engine of the present invention.

FIG. 7 is a valve opening timing change diagram in a case where the opening timing of an exhaust valve is changed depending on the engine water temperature in the compression ignition type internal combustion engine of the present invention.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 Compression ignition type gasoline engine (compression ignition type internal combustion engine) 4 Cylinder 5 Piston 6 Intake port 7 Exhaust port 8 Intake valve 9 Exhaust valve 12 Opening / closing timing variable means 14 Combustion chamber 15 Port injection valve 16 In-cylinder injection valve

──────────────────────────────────────────────────続 き Continued on the front page (51) Int.Cl. ⁶ Identification code FI F02D 41/40 F02D 41/40 AC F02M 63/00 F02M 63/00 P

Claims (2)

[Claims] 1. A cylinder, a piston reciprocating in the cylinder, a combustion chamber defined by the cylinder and the piston, and an intake valve for opening and closing an intake port for supplying intake air to the cylinder. An exhaust valve that opens and closes an exhaust port for exhausting gas in the cylinder, and an opening / closing timing variable unit that changes the opening / closing timing of only the intake valve or both the intake valve and the exhaust valve among the intake valve and the exhaust valve;
In a compression ignition type internal combustion engine including a port injection valve that injects fuel into the intake port and an in-cylinder injection valve that directly injects fuel into the combustion chamber, when fuel is injected only from the port injection valve, The maximum injection amount at which knocking does not occur is set as the reference injection amount, and the corresponding load amount is set as the reference load amount. The opening / closing timing variable means is controlled such that the injection is performed from the port injection valve during the intake stroke after the top dead center and the valve opening overlap period of the intake valve and the exhaust valve increases as the load decreases. Compression ignition type internal combustion engine. 2. A cylinder, a piston reciprocating in the cylinder, a combustion chamber defined by the cylinder and the piston, and an intake valve for opening and closing an intake port for supplying intake air to the cylinder. An exhaust valve for opening and closing an exhaust port for exhausting gas in the cylinder, a port injection valve for injecting fuel into the intake port, and an in-cylinder injection valve for directly injecting fuel into the combustion chamber. In the compression ignition type internal combustion engine, a maximum injection amount at which knocking does not occur when fuel is injected only from the port injection valve is set as a reference injection amount, and a load amount corresponding to the maximum injection amount is set as a reference load amount. In the load range exceeding the required amount, the fuel for the reference injection amount out of the required injection amount is injected from the port injection valve, and the required injection amount is insufficient in the reference injection amount. Fuel compression ignition type internal combustion engine, characterized in that it is injected from in-cylinder injector in the compression top dead center.