JPH05248277A - Cylinder injection spark ignition engine - Google Patents

Abstract

PURPOSE:To improve combustion at low load time without increasing an output loss by delaying the closing timing of both intake and exhaust valves with an overlap thereof left as held constant, when a load of an internal combustion engine is a predetermined value or less. CONSTITUTION:An engine 1 is provided with fuel injection valves 5 for injecting fuel directly into cylinders. Fuel in a fuel tank 9 is supplied through a high pressure reservoir tank 6 by low and high pressure fuel pumps 10, 7. An ECU (electronic control unit) 20 controls the fuel injection valves 5 and spark plugs 16 based on each detection signal from a crank angle sensor 22 and an accelerator opening sensor 24. Here in a cam shaft for driving intake and exhaust valves, a variable valve timing device 30 is additionally provided. In the ECU20, when a load of the engine 1 is a predetermined value or less, an overlap of the intake and exhaust valves is left as held constant, to delay the closing timing of both the intake and exhaust valves.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a cylinder injection type spark ignition engine.

[0002]

2. Description of the Related Art A fuel injection valve for directly injecting fuel into a cylinder is provided, and at least a part of the required fuel injection amount is injected during a cylinder compression stroke during a relatively low load operation in which the fuel injection amount is small and a spark plug is provided. A cylinder direct injection type spark ignition engine is known in which a lean air-fuel mixture combustion is performed by stratifying an air-fuel mixture in the vicinity.

An example of this type of internal combustion engine is proposed by the applicant of the present application in Japanese Patent Laid-Open No. 169834/1990. In the in-cylinder direct injection spark ignition engine proposed in the above publication, when the required fuel injection amount is smaller than the predetermined first injection amount, the entire required fuel injection amount is injected during the cylinder compression stroke to obtain the required fuel injection amount. The injection amount is larger than the first injection amount,
If it is smaller than the second predetermined injection amount, part of the required fuel injection amount is injected during the compression stroke, and the remaining fuel is injected during the intake stroke in advance.

By injecting the fuel in this manner, the fuel injected during the intake stroke forms a homogeneous lean mixture capable of flame propagation throughout the cylinder, and the fuel injected during the compression stroke spark plugs. Since it is stratified in the vicinity and forms a relatively rich air-fuel mixture that can be ignited, the ignition of the air-fuel mixture and the propagation of flame are improved, and the combustion state is improved. In addition, when the required fuel injection amount is relatively large and the total amount of fuel is injected during the compression stroke, there is a problem of smoke generation and output reduction caused by locally forming a rich mixture near the spark plug. Will be resolved.

When the required fuel amount further increases and becomes equal to or larger than the second injection amount and it becomes possible to form a uniform air-fuel mixture having a concentration that can be ignited by the spark plug only in the intake stroke injection, the compression stroke The injection is stopped so that the entire required fuel injection amount is injected during the intake stroke. (Alternatively, also in this case, a part of the fuel injection amount may be injected during the compression stroke.) By performing the divided fuel injection as described above, stable ignition is achieved in the low load region where the fuel injection amount is small. With getting
In a region where the fuel injection amount is large, it is possible to improve the air utilization rate and obtain good combustion.

However, if the stratified mixture combustion is mainly performed during the low load operation as described above, the temperature of the combustion chamber wall becomes relatively low because the gas temperature does not rise sufficiently even by the combustion around the stratified mixture region. Tend. For this reason, the air-fuel mixture is not sufficiently heated in the combustion chamber, and the air-fuel mixture temperature before ignition is low. Therefore, the combustion speed is low and the combustion gas temperature in the cylinder is low. As a result, the temperature of the combustion chamber wall does not rise during stratified mixture combustion, and combustion is likely to become unstable due to a decrease in combustion speed.

In order to solve this problem, the applicant of the present invention has
Japanese Patent Application No. 2-311656 has already proposed an engine in which the overlap of intake and exhaust valves is increased when the engine is operating under a low load than when it is operating under a high load. That is, the engine proposed in Japanese Patent Application No. 2-311656 is provided with a variable valve timing device for the intake valve, and when the engine load is low, the valve timing of the exhaust valve is kept constant and the intake valve valve timing is advanced. Exhaust valve overlap is increased.

In this way, by advancing the valve timing of the intake valve and advancing the opening timing of the intake valve, the burnt gas in the cylinder is back-flowed to the intake port when the intake valve is opened, and then in the intake stroke. The amount of residual burned gas in the cylinder can be increased by reinjecting the burned gas into the cylinder. The high-temperature burned gas remaining in the cylinder mixes with the fresh air drawn in to raise the temperature of the air-fuel mixture, so the in-cylinder gas temperature before ignition rises, and the above-mentioned during stratified combustion due to the decrease in gas temperature. The problem can be resolved.

[0009]

However, when the opening timing of the intake valve is advanced to increase the valve overlap at the time of low load, there is a problem that the closing timing of the intake valve is correspondingly advanced. In an engine that performs stratified mixture combustion at low load, the amount of intake air is usually reduced in order to reduce pumping loss, so the amount of intake air is large compared to the amount of fuel injection, and most of the intake air is Is not involved in combustion. Therefore, as the air taken into the cylinder increases, the work for compressing the air that is not involved in combustion increases in the compression stroke. Therefore, when the load is low, it is preferable to set the closing timing of the intake valve to a timing relatively late in the compression stroke so that part of the air taken into the cylinder in the intake stroke is pushed out to the intake port in the compression stroke. .. This is because the amount of intake air remaining in the cylinder is reduced, and the work for compressing the air that is not involved in combustion can be reduced.

However, as described above, if the closing timing of the intake valve is advanced when the load is low, the amount of intake air remaining in the cylinder is rather increased. For this reason, when the valve timing of the intake valve is advanced to increase the overlap, the combustion state of the stratified mixture can be improved by increasing the amount of residual burnt gas, but the intake air compression that does not contribute to combustion is compressed. There is a possibility that the work required for the engine will increase, and the engine output will decrease and the fuel consumption will increase.

In view of the above, it is an object of the present invention to improve combustion at a low load of a direct injection spark ignition engine without increasing output loss.

[0012]

According to the present invention, a fuel injection valve capable of directly injecting fuel into a cylinder is provided, and the fuel injected into the cylinder is brought to the vicinity of the spark plug in an operation in which the engine load is below a predetermined value. In an in-cylinder injection spark ignition engine, in which the injected fuel is stratified and the fuel injected into the cylinders is evenly distributed and homogeneously burned when the engine load exceeds a predetermined value, the intake valve and the exhaust valve In-cylinder injection type characterized by providing a means for changing the timing and delaying the closing timing of both the intake and exhaust valves while keeping the overlap of the intake and exhaust valves constant when the engine load is below a predetermined value A spark ignition engine is provided.

[0013]

The operation of the present invention will be described with reference to the valve timing chart of FIG. FIG. 1 shows opening and closing timings of the intake and exhaust valves.
(A) shows the valve timing when the engine load is low and medium.
(B) shows the valve timing under high load. Further, TDC and BDC in the figure respectively indicate the top dead center and the bottom dead center of the piston, EO and EC indicate the opening timing and closing timing of the exhaust valve, and IO and IC indicate the opening timing and closing valve of the intake valve. It indicates the time. The hatched portion in the figure shows the timing (valve overlap) when both the intake and exhaust valves are open. 1 (a) and 1 (b), the exhaust valve is opened before the piston reaches BDC in the expansion stroke, the high pressure and high temperature burnt gas in the cylinder is discharged to the exhaust system, and the piston moves in the exhaust stroke. The valve is closed after reaching TDC.

According to the present invention, as shown in FIG. 1 (a), at low and medium loads, the exhaust valve is closed without changing the valve overlap of the intake and exhaust valves under high load (FIG. 1 (b)). Since the timing (EC) is delayed, the exhaust valve continues to open after the piston enters the intake stroke. Therefore, the so-called internal EGR amount in which a part of the burnt gas flowing out to the exhaust port in the exhaust stroke flows into the cylinder from the exhaust port again in the intake stroke increases, and the residual burned gas amount in the cylinder increases.

As a result, at low and medium loads (FIG. 1 (a)), the amount of residual burned gas can be increased without changing the valve overlap, the gas temperature in the cylinder rises, and the combustion of the stratified mixture is performed. The condition is improved. Further, in the present invention, at the time of low and medium load, the closing timing of the intake valve (IC) together with the closing timing of the exhaust valve.
Is delayed (FIG. 1 (a)).

Therefore, when the load is low and medium, the period during which the intake valve continues to open after the piston enters the compression stroke is longer than when the load is high (FIG. 1B). Therefore, of the fresh air that has flowed into the cylinder during the intake stroke, the amount of fresh air that is pushed out again into the intake port during the compression stroke increases, and when the intake valve is closed (I
In C), the amount of air remaining in the cylinder decreases. In an engine that performs stratified mixture combustion at low and medium loads, since the in-cylinder air excess ratio is originally extremely large, combustion does not deteriorate due to the above-mentioned decrease in air amount, which reduces the in-cylinder air amount in the compression stroke. However, the extra compression work can be reduced.

[0017]

FIG. 2 shows the construction of an embodiment of a 4-cylinder gasoline engine to which the present invention is applied. In the figure, 1 is an engine body, 2 is an intake pipe, 3 is an air cleaner, and 4 is a surge tank. In this embodiment, the intake pipe is not provided with a throttle valve and the intake air amount is not controlled. Although not shown, the engine of this embodiment has a double overhead camshaft (DOHC) type valve operating mechanism for driving intake and exhaust valves by different camshafts, and the camshaft will be described later. A variable valve timing device (VVT) 30 is provided so that the opening and closing timings of the intake valve and the exhaust valve can be simultaneously changed by the same amount.

Further, the engine of this embodiment is provided with a fuel injection valve 5 for directly injecting fuel into each cylinder. The fuel injection valve 5 uses a piezo-piezoelectric element having a high operating speed in order to perform the divided injection of fuel during one stroke of each cylinder. 6 is a high pressure reservoir tank for supplying high pressure fuel to the fuel injection valve 5 through each branch pipe 14, and 7 is a high pressure conduit 8
A high-pressure fuel pump capable of controlling the discharge pressure for pumping high-pressure fuel to the reservoir tank 6 via a fuel tank, 9 is a fuel tank, and 10 is a low-pressure fuel for supplying fuel from the fuel tank 9 to the high-pressure fuel pump 7 via a conduit 11. Fuel pumps are shown respectively.
The discharge side of the low-pressure fuel pump 10 has a piezoelectric element cooling introduction pipe 12 for cooling the piezoelectric element of each fuel injection valve 5.
Connected to. The piezoelectric element cooling fuel oil return pipe 13 is connected to the fuel tank 9, and the fuel flowing through the piezoelectric element cooling introduction pipe 12 is collected in the fuel tank 9 via the return pipe 13.

The electronic control unit (ECU) 20 is composed of a digital computer and controls the engine 1 in various ways. Due to these controls, the ECU 20 receives an output pulse proportional to the engine speed Ne from the crank angle sensor 22, and an accelerator pedal (not shown) manipulated variable (accelerator opening) θ _A from the accelerator opening sensor 24. The corresponding output signal is input.

The ECU 20 is connected to an ignition plug 16 provided in each cylinder via an ignition circuit (not shown) to control the ignition timing of each cylinder, and to each fuel injection valve 5 via a drive circuit (not shown). It is connected and controls the fuel injection timing and the fuel injection amount. Similarly, the VVT 30 is also connected to the ECU 20 via a drive circuit (not shown),
The opening and closing timings of the intake and exhaust valves are changed according to the signal from the ECU 20.

As described above, the engine of this embodiment does not control the intake air amount by the throttle valve, and the engine load changes by changing the fuel injection amount according to the accelerator angle θ _A and the engine speed Ne. Controlled. Therefore, since a large amount of intake air is always supplied to the combustion chamber, the air-fuel ratio is leaner than that of a normal engine, and if the injected fuel is diffused evenly into the cylinder, especially at low load, the spark plug Will not be able to ignite. Therefore, in this embodiment, when the engine load is low, the fuel injection timing is delayed so that fuel is injected in the latter half of the cylinder compression stroke. FIG. 3 is a longitudinal sectional view of the engine showing a diffusion pattern of the injected fuel when the fuel is injected in the latter half of the compression stroke.

In the figure, 60 is a cylinder block, and 61
Is a cylinder head, 62 is a piston, 63 is a piston 6
2 is a substantially cylindrical recess formed on the top surface of the piston 2, 64 is the piston 6
2 shows cylinder chambers formed between the top surface and the inner wall surface of the cylinder head 61, respectively. The spark plug 16 faces the cylinder chamber 64 and is attached to a substantially central portion of the cylinder head 61.
An intake port and an exhaust port are formed in the cylinder head 61, and an intake valve and an exhaust valve are arranged at openings of the intake port and the exhaust port into the cylinder chamber 64, respectively. This is not shown in the drawing because the intake and exhaust valves are both closed.
The fuel injection valve 5 is a swirl type fuel injection valve, and injects a fuel atomized fuel having a wide spread angle and a weak penetration force. The fuel injection valve 5 is arranged diagonally downward, is arranged at the top of the cylinder chamber 64, and is arranged so as to inject fuel toward the vicinity of the spark plug 65. Further, the fuel injection direction and the fuel injection timing of the fuel injection valve 5 are determined so that the injected fuel is directed to the recess 63 formed at the top of the piston 62. As shown in the figure, the fuel injected in the latter stage of the compression stroke is in the cylinder chamber 64 because the pressure in the cylinder chamber 64 is rising, the penetration force is weak, and the air flow in the cylinder chamber 64 is also low. It does not diffuse, but is concentrated in the region K near the spark plug 16 and stratified. Since the fuel distribution in this region K is also non-uniform and changes from the rich air-fuel mixture layer to the air layer, the air-fuel mixture layer that is most likely to ignite exists in this region K. Therefore, even if the fuel injection amount is small, ignition by the spark plug 16 is possible.

Next, FIGS. 4A to 4D show the fuel injection in the load region higher than the medium load. When the load is equal to or higher than the medium load, the fuel is divided into two parts, that is, the initial stage of the intake stroke (FIG. 4A) and the latter stage of the compression stroke (FIG. 4C). This is because the fuel injection amount also increases in the region of medium load or higher, so that if the total amount of fuel is injected in the latter half of the compression stroke, the air-fuel mixture region formed near the spark plug 16 becomes totally rich, and in this region This is because the amount of air is insufficient and incomplete combustion occurs, which may cause the generation of exhaust smoke and insufficient output.

FIG. 4 (a) shows the first fuel injection at the beginning of the intake stroke. In this state, the intake valve 66 is open, and fresh air is flowing into the cylinder chamber 64 from the intake port. Therefore, the fuel injected from the fuel injection valve 5 is diffused by the turbulence R caused by the new airflow flowing from the intake port, and the cylinder chamber 6 is reached during the intake stroke to the compression stroke.
A uniform air-fuel mixture P is formed in 4 (FIG. 4 (b)). The air-fuel ratio of the air-fuel mixture P is considerably lean, and it is difficult to ignite it directly by the spark plug 16, but once ignited, the air-fuel ratio is set to a value at which the ignition flame can propagate. Next, after the intake valve 66 is closed, the second fuel injection is performed in the latter half of the compression stroke (FIG. 4 (c)).
The fuel injected in the latter part of the compression stroke forms a mixture region K containing a mixture having a concentration capable of being ignited in the vicinity of the spark plug 16 as in the case of FIG. Therefore, when ignition is performed by the spark plug 16, combustion progresses mainly in the air-fuel mixture region K, and the flame propagates to the air-fuel mixture P sequentially from the periphery thereof and the combustion progresses. Since the fuel injection amount is not so large in the medium load range, if all the fuel is injected during the intake stroke and diffused uniformly in the cylinder chamber 64, ignition by the spark plug 16 becomes difficult. By injecting fuel into the intake stroke and the compression stroke in a divided manner to form an air-fuel mixture having a concentration higher than that capable of ignition flame propagation and a stratified air-fuel mixture having a concentration capable of being ignited by the spark plug 16. , It is possible to ignite and burn a lean mixture as a whole. Further, in the present embodiment, the load increases, the fuel injection amount increases, and the cylinder chamber 6
When the amount that can form a uniform air-fuel mixture with a concentration capable of being ignited by the spark plug 16 in 4 is reached, the injection during the compression stroke is stopped and the entire amount of fuel is injected at the beginning of the intake stroke.

FIG. 5 is a diagram showing the relationship between the fuel injection amount Q _S during the intake stroke and the fuel injection Q _C during the compression stroke. The horizontal axis represents the total injection amount Q _T = Q _S + Q _C and the vertical axis represents the total. The distribution of the injection quantity Q _T between the compression stroke injection quantity Q _c and the intake stroke injection quantity Q _S is shown. As can be seen from the figure, when the total injection amount Q _T is Q ₁ or less (section I), Q _S is set to zero and the entire amount is injected during the compression stroke. When Q _T is Q ₁ or more and Q ₂ or less (section II), Q _C is reduced to Q _C1 which is sufficient to form an ignitable stratified mixture near the spark plug, and other amounts are It is injected during the intake stroke. When Q _T becomes equal to or greater than Q ₂ (section III), the compression stroke injection amount Q _C is set to zero and all fuel is injected during the intake stroke. Q ₂ is a fuel injection amount sufficient to form a uniform mixture in the cylinder chamber 64 that can be ignited by the spark plug 16.

In this embodiment, the total fuel injection amount is Q ₂ or less,
In particular, in the region of less than Q ₁ , the combustion in the cylinder chamber is mainly the stratified mixture combustion. Therefore, as described above, the cylinder wall temperature does not rise and the air-fuel mixture temperature also drops, so that combustion after ignition is likely to be unstable. In this embodiment, the open / close timings of both the intake valve and the exhaust valve are changed according to the load (total fuel injection amount) to increase the burnt gas amount in the cylinder during stratified mixture combustion and stabilize combustion. It prevents the increase of compression work.

That is, in this embodiment, the total fuel injection amount Q
_{When T} is equal to or less than Q ₂ (FIG. 5, sections I and II), that is, when stratified mixture combustion is performed, the opening and closing timings of the intake valve and the exhaust valve are delayed by the same amount to the state of FIG. 1 (a). To be Further, when the total fuel injection amount Q _T is Q ₂ or more (FIG. 5, section III), that is, when a homogeneous mixture combustion is performed, the intake / exhaust valve opening / closing timing is advanced by the same amount, and FIG. To be in the state of. As a result, at the time of low and medium load (Fig. 1 (a)), the amount of residual burned gas in the cylinder is increased, and at the same time, the amount of air compressed in the compression stroke is reduced, so that both combustion stability and output loss are reduced. Can be obtained. Further, at the time of high load (Fig. 1 (a)), the amount of residual burnt gas is reduced by advancing the closing timing of the exhaust valve, and further advancing the closing timing of the intake valve, so that it is introduced into the cylinder in the compression stroke. Since the amount of remaining fresh air increases, the intake volume efficiency improves,
Large output can be secured.

Next, FIG. 6 shows a variable valve timing device 30 used in this embodiment. In this embodiment, as the variable valve timing device 30, the one described in JP-A-58-135310 is used, but another type of variable valve timing device can also be used. 6 to 8 below
The variable valve timing device disclosed in JP-A-58-135310 will be described with reference to FIG. In the present embodiment, the intake valve cam shaft is provided with the variable valve timing device, and the valve timing of the intake valve is changed by changing the rotation phase of the intake cam shaft with respect to the crank shaft. Further, although not shown, in the engine of this embodiment, the exhaust camshaft is driven synchronously from the intake valve via a gear, and the exhaust camshaft is simultaneously driven by changing the rotational phase of the intake camshaft with respect to the crankshaft. The rotation phase of is also changed by the same amount. That is, the variable valve timing device of this embodiment changes the opening / closing timings of the intake valve and the exhaust valve by the same amount at the same time.

FIG. 6 is a sectional view showing the structure of the variable valve timing device. In FIG. 6, 201 is a timing gear for driving the cam shaft, which is rotationally driven from the crank shaft by a belt (not shown). Further, 202 is a cam shaft which drives an intake valve, and the timing gear 201 and the cam shaft 202 are respectively an outer sleeve 203 and an inner sleeve 204.
It is fixed to and. The outer sleeves 203 and 204 are attached so as to be rotatable relative to each other, and the outer sleeves 203 and 204 are mounted on the outer circumferences of the respective sleeves 1 at axially symmetrical positions as shown in FIG.
A pair of slits 205, 205B and 206, 206B are provided. As shown in FIG. 8, the slit 205 on the outer sleeve and the slit 206 on the inner sleeve are formed so as to be inclined in directions opposite to each other with respect to the axial direction of the cam shaft 202, and are formed at the intersections of the slits 205 and 206. Are roller bearings 207 that can rotate independently of each other
And 208 are provided to contact one of the inner walls of the slits 205 and 206, respectively. Here, the slits 205B and 20 at positions axially symmetrical to those shown in FIG.
Similarly, on the inner wall of 6B, roller bearings 207B and 20
8B is in contact with the inner wall 205, 20 which is in contact with
6 is the opposite of FIG. Therefore, by simultaneously moving the roller bearings 207, 208 and 207B, 208B back and forth in the axial direction of the cam shaft, the outer sleeve 203 and the inner sleeve 2 do not cause backlash.
04 can be rotated relative to each other. As described above, since the outer sleeve 203 and the inner sleeve 204 are integrally fixed to the timing gear 201 and the cam shaft 202, respectively, the timing gear 20 is rotated by the relative rotation of the outer sleeve 203 and the inner sleeve 204.
1 changes the phase between the camshaft 202 and the camshaft 202, so that the phase angle of the valve opening / closing timing with respect to the crankshaft can be changed.

Next, a mechanism for moving the roller bearings 207 (B) and 208 (B) in the axial direction of the cam shaft will be described. As shown in FIGS. 6 and 7, the roller bearings 207 (B) and 208 (B) are rotatably supported by a bearing shaft 211 which is inserted and supported in the diameter portion through hole 210 of the slider 209. Two
09 is rotatably supported by a non-rotating drive sleeve 212, which is movable in the axial direction but is fixed in the rotating direction, via a bearing 213. The outer race and the inner race of the bearing 213 are respectively sliders 209.
And the drive sleeve 212. Therefore, if the drive sleeve 212 moves in the axial direction while it is not rotating, the slider 209 and the bearing shaft 211 will move to the outer sleeve 2.
03, the inner sleeve 204 and the inner sleeve 204 can move in the axial direction while being rotatably held around the slider 209. Further, a screw thread is provided on the inner surface of the drive sleeve, and is screwed with a screw thread of the output shaft 216 of the step motor 215 fixed to the housing 214. The drive sleeve 212 is an axial spline 2 provided on the housing.
Since the output shaft 216 is fixed to the rotation by the step motor 17 when the output shaft 216 is rotated by the step motor,
12 moves in the axial direction. That is, the step motor 21
When the slider 209 and the roller bearings 207 (B) and 208 (B) are rotated by rotating the motor 5,
It is possible to change the phase of the camshaft 202 with respect to the crankshaft by moving the outer sleeve 203 and the inner sleeve 204 relative to each other by axially moving by a distance corresponding to the rotation of the shaft 15. As is clear from the above description, in the present embodiment, the valve opening timing and the valve closing timing of the intake valve and the exhaust valve simultaneously change by the same phase, and the valve opening period is kept constant.

In the present embodiment, as shown in FIG. 9, the retard angle of the intake / exhaust valve opening / closing timing is set in the region where the engine is operated at a relatively low load (low torque), that is, the region where fuel injection is performed in the compression stroke. By doing so, the combustion state is improved and the compression work is reduced during stratified mixture combustion. Next, the control operation by the ECU 20 of the present embodiment will be described using the flowchart of FIG.

FIG. 10 shows a fuel injection control routine.
The routine is executed by interruption at every constant crank angle.
Be done. In the figure, first in step 100
Number of revolutions Ne and accelerator opening θ_AIs read. Ste
In Map 102, from map 1 (FIG. 11), Ne and θ
_AThe required fuel injection amount Q based on_TIs calculated. 11
Q_TIs θ _AIncreases with increasing N
e has a maximum value near 4000 rpm.

At step 104, the required fuel injection amount Q_T
Is Q₁(See FIG. 5) It is determined whether or not the following. Affirmative
If so, the required fuel injection amount Q in step 106 and thereafter_Tof
The entire quantity is injected in the compression stroke. That is, the step
In compression stroke fuel injection amount Q_CIs the required fuel injection amount Q
_TIs set, and in step 108, the intake stroke fuel injection amount Q
_SIs set to 0. In step 110, the compression stroke fuel injection period
Interval T_CFrom map 2 (Figure 12) to Q_CCalculated based on
Be done. Referring to FIG. 12, T_CIs Q_CAs the
Increase linearly. In step 112, intake stroke fuel
Injection period T_SIs set to 0. Map at step 114
3 (FIG. 13) to the compression stroke fuel injection amount Q _CAnd engine times
The compression stroke fuel injection start timing T based on the rotation speed Ne_CIIs arithmetic
Issued in step 115, the flag θ_RTDSet 1 to
This routine ends. Referring to FIG. 13, T_CIIs pressure
It is indicated by the injection advance angle from top dead center. T_CIIs Ne
And Q_CIs accelerated as the number increases. θ_RTDWill be described later
Is a flag that indicates the intake / exhaust valve opening / closing timing.
R θ_RTDWhen = 1, the intake / exhaust valve opening / closing timing is delayed by a predetermined amount.
Be horned.

When a negative determination is made in step 104, it is then determined in step 116 whether the required fuel injection amount Q _T is equal to or less than Q ₂ (see FIG. 4). When the determination is affirmative, the required fuel injection amount Q _T is divided into the compression stroke and the intake stroke in step 118 and subsequent steps. In step 118, Q _C1 (see FIG. 5) is put into the compression stroke fuel injection amount Q _C. Q _T -Q _C1 is placed in the step 120, the intake stroke fuel injection amount Q _S. That is, the sum of the intake stroke fuel injection amount Q _S and the compression stroke fuel injection amount Q _C is set to the required injection amount Q _T. In steps 122 and 124, map 2 (see FIG.
2) to Q _C, Q _S based on the compression stroke fuel injection period T _C
And the intake stroke fuel injection period T _S are calculated.

Next, at step 126, the map 4 (see FIG.
4) Required injection amount Q_TAnd based on engine speed Ne
Compression stroke fuel injection start timing T_CIIs calculated. 14
See, T_CIIs indicated by the injection advance angle from the compression top dead center
And T_CIIs Ne and Q _TFaster as
To be In step 128, from map 5 (Fig. 15)
Intake stroke fuel injection amount Q_SAnd based on engine speed Ne
Intake stroke fuel injection start timing T_SIIs calculated. Figure 15
See, T_SIIs indicated by the injection advance angle from the compression top dead center
And T_SIIs accelerated as Ne increases
But Q_SIt can't be changed by. This is the intake
Stroke injection is sufficient for the fuel to diffuse and form a mixture
I have a lot of time, so Q_SNeed to be changed depending on the
Because there is no. Step 12 after the above steps are completed
Flag θ at 9_RTDSet 1 to and end the routine
It

When a negative determination is made in step 116, the total amount of the required fuel injection amount Q _T is injected during the intake stroke in step 130 and subsequent steps. That is, at step 130, the intake stroke fuel injection amount Q _S is set to the required fuel injection amount Q _T ,
In step 132, the compression stroke fuel injection amount Q _C is made zero. Further, the compression stroke fuel injection period T _C is also set to 0 in step 134, and in step 136 the map 2 (FIG. 12) to Q
_The intake stroke fuel injection period T _S is calculated based on _S.
Next, at step 138, the intake stroke fuel injection start timing T _SI is calculated from the map 5 (FIG. 15) based on Q _S and Ne, and at step 139 the flag θ _RTD is reset to 0 and the routine ends.

After the above steps are completed, fuel injection is executed by another routine (not shown). Similarly, the ignition timing is set by another routine (not shown) based on the required fuel injection amount Q _T and the engine speed Ne. Next in FIG.
Reference numeral 6 is a routine showing the opening / closing timing control of the intake / exhaust valve. This routine is executed by interruption at every constant crank angle. In the figure, in step 150, it is determined whether or not θ _RTD is 1, and if θ _RTD is 1, the intake valve opening / closing timing is delayed, so the intake valve opening timing θ _I is the predetermined crank angle θ _a. Is set to (step 15
2). When the intake valve opening timing θ _I is retarded to θ _a , the phase of the exhaust cam shaft driven by the intake cam shaft via the gear changes with the crank shaft by the same amount, and the exhaust valve opening / closing timing Is also delayed (FIG. 1 (a)). As a result, the amount of burnt gas remaining in the cylinder is increased, and at the same time, the compression work is reduced.

If θ _RTD is 0 in step 150, the intake valve opening / closing timing is advanced by a predetermined amount, and the intake valve opening timing θ _I is set to the crank angle θ _b (step 154). As a result, the opening / closing timing of the intake / exhaust valve is set to the state shown in FIG. 1 (b), and a high output can be secured by improving the intake volume efficiency. When the setting of θ _I is completed as described above, the value of θ _I is output to the drive circuit of the variable valve timing device in step 156, and the step motor 215 of the variable valve timing device 30 is driven to change the intake valve opening timing.

In the present embodiment, the intake valve opening / closing timing is fixed at two values, the normal value θ _a and the advance value θ _b , but depending on the compression stroke fuel injection amount or the intake stroke fuel injection amount. The intake valve opening / closing timing may be changed continuously. Further, in this embodiment, the variable valve timing device 30 is used.
Is provided on the intake cam shaft and the exhaust cam shaft is driven from the intake cam shaft through a gear to simultaneously change the valve timing of the intake and exhaust valves, but the present invention is not limited to this. For example, when driving a camshaft from a crankshaft using a belt, a chain, etc., a variable valve timing device is installed between the crankshaft and a pulley or sprocket so that the valve timings of intake and exhaust valves can be changed simultaneously. May be.

[0040]

The cylinder injection type spark ignition engine according to the present invention, when the engine load is low, delays the closing timing of both the intake valve and the exhaust valve as compared with the high load, so that the engine is operated at low load. In the above, there is an excellent effect that it is possible to stabilize the stratified mixture combustion and reduce the loss due to the compression work, and to secure a large output during high load operation.

[Brief description of drawings]

FIG. 1 is a diagram illustrating an operation of the present invention by changing a valve timing of an intake / exhaust valve.

FIG. 2 is an overall configuration diagram showing an embodiment of an engine to which the present invention is applied.

FIG. 3 is a diagram showing a mixture state in a cylinder chamber at the time of fuel injection in a compression stroke.

FIG. 4 is a diagram showing a mixture state in a cylinder chamber during split injection.

FIG. 5 is a diagram showing fuel injection amounts of compression stroke injection and intake stroke injection.

FIG. 6 is a diagram showing a structure of a variable valve timing device.

FIG. 7 is a diagram showing a structure of a variable valve timing device.

FIG. 8 is a diagram showing a structure of a variable valve timing device.

FIG. 9 is a diagram showing a relationship between a fuel injection method and valve timing setting.

FIG. 10 is a flowchart showing a fuel injection control operation.

FIG. 11 is a diagram showing a map used for setting a fuel injection amount θ _T.

FIG. 12 is a diagram showing a map used for setting a fuel injection period T _C.

FIG. 13 is a diagram showing a map used for setting fuel injection start timing T _CI .

FIG. 14 is a diagram showing a map used for setting fuel injection start timing T _CI .

FIG. 15 is a diagram showing a map used for setting a fuel injection start timing T _SI .

FIG. 16 is a flowchart showing an intake / exhaust valve timing control operation.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 ... Engine main body 2 ... Intake pipe 3 ... Air cleaner 4 ... Surge tank 5 ... Fuel injection valve 6 ... High pressure reservoir 7 ... High pressure pump 9 ... Fuel tank 10 ... Low pressure pump 16 ... Spark plug 20 ... Electronic control unit 30 ... Variable valve timing apparatus

Claims (1)

[Claims] 1. A fuel injection valve capable of directly injecting fuel into a cylinder is provided, and in an operation in which the engine load is a predetermined value or less, the fuel injected into the cylinder is guided to the vicinity of the spark plug for stratified combustion so that the engine load is predetermined. In an in-cylinder injection spark ignition engine that distributes the fuel injected into the cylinders uniformly in the combustion chamber when operating above the specified value, a means for changing the valve timing of the intake valve and the exhaust valve is provided for the engine load. Is a predetermined value or less, a cylinder injection type spark ignition engine characterized by delaying the closing timing of both intake and exhaust valves while keeping the overlap of intake and exhaust valves constant.