JPH04183945A - Cylinder direct injection type spark ignition engine - Google Patents

Abstract

PURPOSE:To stabilize combustion of stratified air-fuel mixture, increase output in the case of low load operation and improve properties of exhaust gas by arranging a variable valve timing device at least in one of intake and exhaust valves, and increasing overlap of the intake and exhaust valves in the case of the low load operation more than that in the case of high load operation. CONSTITUTION:A variable valve timing device 30 is arranged on a cam shaft to drive an intake valve, and fuel injection valves 5 to inject fuel directly are also arranged in respective cylinders. Since the respective fuel injection valves 5 carry out divided injection of the fuel in one stroke of the respective cylinders, the valves having high operational speed equipped with piezoelectric elements are used therefor, and in the case an engine load is less than a prescribed value, at least a part of requested fuel injection volume according to an engine operational condition is injected into the cylinders so that air-fuel mixture can be formed in the vicinity of sparking plugs 16. In order to carry out such various controls of an engine 1, a ECU 20 to input output signals from a crank angle sensor 22, an accel opening sensor 24 and so on are arranged, and the variable valve timing device 30 is controlled by means of this ECU 20 so that overlap of intake and exhaust valves in the case of low load operation can be increased more than that in the case of high load operation.

Description

[Detailed description of the invention] [Industrial application field] The present invention relates to an in-cylinder direct injection spark ignition engine.

[Conventional technology]

Equipped with a fuel injection valve that directly injects fuel into the cylinder, during relatively low-load operation with a small fuel injection amount, at least a portion of the required fuel injection amount is injected during the cylinder compression stroke to create a mixture near the spark plug. An in-cylinder direct injection spark ignition engine that performs lean mixture combustion by stratifying fuel is known.

An example of this type of internal combustion engine is one proposed in Japanese Patent Laid-Open No. 2-169834 by the applicant of the present invention.

In the in-cylinder direct injection spark ignition engine proposed in the same bulletin,
If the required fuel injection amount is less than the predetermined first injection amount, the entire amount of the required fuel injection amount is injected during the cylinder compression stroke, and if the required fuel injection amount is greater than the first injection amount, and If the required fuel injection amount is smaller than the second predetermined injection amount, a part of the required fuel injection amount is injected during the compression stroke, and the remaining fuel is injected in advance during the intake stroke.

By splitting the fuel injection in this way, the fuel injected during the intake stroke forms a homogeneous lean mixture that allows flame propagation throughout the cylinder, and the fuel injected during the compression stroke forms a stratified mixture near the ignition plug. The fuel mixture forms a relatively rich mixture that can be ignited, which improves the ignition of the mixture and the propagation of flame, resulting in improved combustion conditions. In addition, when the required fuel injection amount is relatively large and the entire amount of fuel is injected during the compression stroke, the problem of smoke generation and output reduction caused by locally forming a rich mixture near the ignition plug can be avoided. resolved.

When the required amount of fuel increases further to reach the second injection tooth side, and it becomes possible to form a homogeneous mixture in the cylinder metal body with a concentration that can be ignited by the ignition plug by only intake stroke injection, the compression stroke Injection is stopped so that the entire required fuel injection amount is injected during the intake stroke. (Alternatively, in this case as well, part of the fuel injection amount may be injected during the compression stroke.) By performing split fuel injection as described above, stable ignition can be achieved in low load regions where the fuel injection amount is small. At the same time, it is possible to improve the air utilization rate and obtain good combustion in areas where the amount of fuel injection is large.

[Problem to be solved by the invention]

The direct-injection spark ignition engine disclosed in JP-A-2-169834 is generally capable of combustion with a lean mixture, but combustion becomes unstable during low-load operation with a small amount of fuel injection. There are cases. That is, during low-load operation, stratified air-fuel mixture combustion is mainly performed, so the gas temperature around the air-fuel mixture stratified region does not rise sufficiently due to combustion, and the combustion chamber wall temperature tends to become relatively low. For this reason, the temperature of the intake air is not sufficiently raised in the combustion chamber, and the temperature of the air-fuel mixture before ignition becomes low. A decrease in the mixture temperature delays the vaporization of the fuel, leading to a decrease in the combustion rate of the mixture and lowering the combustion gas temperature in the cylinder. Therefore, during combustion of a stratified mixture, the combustion chamber wall temperature does not increase and the combustion rate decreases. This decrease may cause combustion to become unstable, resulting in increased engine vibration, decreased output, and worsened exhaust gas properties.

In order to solve the above problem, it is possible to reduce the amount of intake air at low loads to reduce the amount of combustion gas and raise the cylinder temperature, but this increases the pumping loss due to intake throttling, resulting in a decrease in engine output. This is not preferable because it increases fuel consumption.

In view of the above, an object of the present invention is to improve combustion during low load in a direct injection spark ignition engine.

[Means to solve the problem]

According to the present invention, a fuel injection valve capable of directly injecting fuel into the cylinder of an engine is provided, and when the engine load is below a predetermined value, at least a part of the required fuel injection amount according to the engine operating state is injected into the cylinder. In a direct injection spark ignition engine that forms a mixture near the ignition plug, a variable valve timing device is installed on at least one of the intake valve and exhaust valve, so that when the engine is running at low load, the timing is lower than when the engine is running at high load. There is also provided an in-cylinder direct injection spark ignition engine characterized in that the overlap between intake and exhaust valves is increased.

5. Effect] When the engine is under low load, the overlap between the intake and exhaust valves increases, which increases the amount of high-temperature burnt gas remaining in the combustion chamber. This residual burnt gas mixes with fresh air during the intake stroke, increasing the gas temperature in the cylinder and keeping the mixture temperature high before ignition, promoting fuel vaporization and increasing the combustion rate, resulting in stratified combustion. Combustion is sometimes stabilized.

〔Example〕

FIG. 1 shows the configuration of an embodiment of a four-cylinder gasoline engine to which the present invention is applied. In the figure, 1 is the engine body, 2
3 indicates the intake pipe, 3 indicates the air cleaner, and 4 indicates the surge tank. In this embodiment, a throttle valve is not provided in the intake pipe, and the amount of intake air is not controlled. Although not shown, the engine of this embodiment has a double overhead camshaft (DOHC) type in which the intake and exhaust valves are driven by independent camshafts, and the camshafts that drive the intake valves are equipped with variable valve timing, which will be described later. Equipment (VVT
) 30 is provided so that the opening/closing timing of the intake valve can be changed.

Furthermore, the engine of this embodiment is provided with a fuel injection valve 5 that directly injects fuel into each cylinder.

The fuel injection valve 5 uses a piezoelectric element with a high operating speed to perform split injection of fuel during one stroke of each cylinder. 6 is a fuel injection valve 5 through each branch pipe 14.
7 is a high-pressure fuel pump whose discharge pressure can be controlled for supplying high-pressure fuel to the reservoir tank 6 through a high-pressure conduit 8; 9 is a fuel tank; 10 is a high-pressure fuel pump through a conduit 11; A low-pressure fuel pump that supplies fuel from a fuel tank 9 to a high-pressure fuel pump 7 is shown in FIG.

The discharge side of the low-pressure fuel pump 10 includes 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 recovered into the fuel tank 9 via this conduit 13 .

The electronic control unit (EC1l) 20 is composed of a digital computer and performs various controls on the engine 1.

For these controls, [EClI 20 receives an output pulse proportional to the engine speed Ne from the crank angle sensor 22,
Further, an output signal corresponding to the operation amount (accelerator opening) θ9 of an accelerator pedal (not shown) is manually generated from the accelerator opening sensor 24.

Further, the EC[I 20 is connected to the ignition plug 16 provided in each cylinder via an ignition circuit (not shown) to control the ignition timing of each cylinder, and is also connected to each fuel injection valve 5 via a drive circuit (not shown). and controls the fuel injection timing and fuel injection amount. Similarly, VVT30 is also connected to [EClI20] via a drive circuit (not shown).
The opening/closing timing of the intake valve is changed according to the signal from U20.

As mentioned above, the engine of this embodiment does not control the intake air amount using the throttle valve, and the engine load is controlled by changing the fuel injection amount according to the accelerator angle θ4 and the engine rotation speed Ne. . Therefore, a large amount of intake air is always supplied into the combustion chamber, so the air-fuel ratio is leaner than that of a normal engine.Especially at low loads, if the injected fuel is uniformly spread within the cylinder, the spark plug ignition becomes impossible. Therefore, in this embodiment, when the engine load is low, the fuel injection timing is delayed so that the fuel injection is performed in the latter half of the cylinder compression stroke. FIG. 2 is a longitudinal sectional view of the engine showing a diffusion pattern of injected fuel when fuel is injected in the latter half of the compression stroke.

In the figure, 60 is a cylinder block, 61 is a cylinder head, 62 is a piston, 63 is a substantially cylindrical recess formed on the top surface of the piston 62, and 64 is a cylinder formed between the top surface of the piston 62 and the inner wall surface of the cylinder head 61. Each room is shown. The ignition plug 16 is attached to a substantially central portion of the cylinder head 61 facing the cylinder chamber 64. An intake port and an exhaust port are formed in the cylinder head 61, and the cylinder chamber 64 of these intake ports and exhaust ports is formed.
An intake valve and an exhaust valve are arranged at the openings to the inside, but they are not shown in the drawing because the figure shows the latter half of the compression stroke and both the intake and exhaust valves are closed. The fuel injection valve 5 is a swirl type fuel injection valve, and injects fuel in the form of a spray with a large spread angle and a weak penetration force. The fuel injection valve 5 is arranged at the top of the cylinder chamber 64 so as to face obliquely downward, and is arranged so as to inject fuel toward the vicinity of the ignition plug 65 . Further, the fuel injection direction and fuel injection timing of the fuel injection valve 5 are determined so that the injected fuel is directed toward the recess 63 formed at the top of the piston 62. As shown in the figure, the pressure in the cylinder chamber 64 of the fuel injected in the latter half of the compression stroke has increased by 1, the penetration force is weak, and the flow of air within the cylinder chamber 64 has also decreased.
It does not diffuse into the interior, but concentrates in the area K near the spark plug 16 and forms a stratification. The fuel distribution within this region is also uneven, varying from a rich mixture layer to an air layer, so the mixture layer that is most likely to ignite exists within this region. Therefore, even when the fuel injection amount is small, ignition by the ignition plug 16 is possible.

Next, FIGS. 3(a) to 3(d) show fuel injection in a load range higher than medium load. At medium or higher loads, fuel is distributed at the beginning of the suction stroke (Fig. 3 (a)) and at the end of the compression stroke (Fig. 3(a)).
Split injection is performed in two parts (Fig. 3(C)). This is because the amount of fuel injection also increases in the region of medium load or higher, so if the entire amount of fuel is injected in the latter half of the compression stroke, the ignition plug 16
This is because the air-fuel mixture region formed in the vicinity becomes overrich as a whole, and the amount of fuel air becomes insufficient in this region, resulting in incomplete combustion, which may result in exhaust smoke or insufficient output.

FIG. 3(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 fresh air flowing from the intake port, and a uniform air-fuel mixture P is formed in the cylinder chamber 64 from the intake stroke to the compression stroke. Figure 3(b)). The air-fuel ratio of this air-fuel mixture P is quite lean, making it difficult to ignite it directly with the ignition plug 16, but once ignited, the air-fuel ratio is at least high enough to allow the ignition flame to propagate. Next, after the intake valve 6G closes, the second half of the compression stroke (third
In Figure (C)), the second fuel injection is performed. The fuel injected in the latter half of the compression stroke forms an air-fuel mixture region K near the spark plug 16 containing an air-fuel mixture with an ignitable concentration, as in the case of FIG.

Therefore, when ignition is performed by the ignition plug 16, combustion progresses centering around the air-fuel mixture region K, and flame propagates sequentially to the air-fuel mixture P from the periphery, and combustion progresses. Since the amount of fuel injected is not very large in a medium load region, if all the fuel is injected during the intake stroke and diffused uniformly within the cylinder chamber 64, it will be difficult for the ignition plug 16 to ignite. The fuel is injected separately into an intake stroke and a compression stroke, and the mixture with a concentration higher than that which allows ignition and flame propagation is injected into the spark plug 1.
By forming a stratified mixture with a concentration that can be ignited by 6, it is possible to ignite and burn an overall lean mixture. In addition, in this embodiment, when the load increases and the fuel injection amount increases and reaches an amount that can form an air-fuel mixture with a uniform concentration that can be ignited by the spark plug 16 in the cylinder chamber 64, during the compression stroke. injection is stopped, and the entire amount of fuel is injected at the beginning of the intake stroke.

Figure 4 shows the fuel injection amount QS during the intake stroke and the fuel injection amount Q during the compression stroke. The horizontal axis is the total injection amount Q, -Q, and 10Q. The vertical axis is the compression stroke injection amount Q of the total injection amount QT. and suction stroke injection amount Q.

It shows the distribution to. As shown in the figure, when the total injection amount Q is less than or equal to Q1 (section I), it is Q.

The entire amount is injected during the compression stroke.

Also, if Oo is greater than or equal to Q1 and less than or equal to 2 (section ■), Q. is reduced to a QCI sufficient to form an ignitable stratified mixture near the spark plug, and the remaining amount is injected during the intake stroke. Also, if ○ becomes Q2 or higher (
In section (3), the compression stroke injection amount Qc is set to zero, and all the fuel is injected during the intake stroke. Q2 is a fuel injection amount sufficient to form a homogeneous air-fuel mixture in the cylinder chamber 64 that can be ignited by the ignition plug 16.

In this embodiment, in a region where the total fuel injection amount is less than Q2, particularly less than Ql, the combustion in the cylinder chamber is mainly stratified temperature amixture combustion. For this reason, as described above, the cylinder wall temperature does not rise and the air-fuel mixture temperature also becomes low, which tends to make combustion after ignition unstable. In this embodiment, this problem is solved by increasing the amount of burnt gas remaining in the shilling chamber by changing the opening and closing timing of the intake valve according to the load (total fuel injection amount).

Figures 5(a) and 5(b) are diagrams showing the hull swimming of the intake and exhaust valves, where TDC and BDC in the figures indicate the top dead center and bottom dead center of the piston, respectively, and E, O, and EC indicate the exhaust valve. I[] indicates the opening timing and closing timing of the intake valve, and IC indicates the opening timing and closing timing of the intake valve. The shaded area in the figure indicates the period when both the intake and exhaust valves are open (valve overlap). The exhaust valve opens before the piston reaches BDC in the expansion stroke, discharges high-pressure, high-temperature burnt gas in the cylinder to the exhaust system, and closes after the piston reaches TDC in the exhaust stroke. On the other hand, the intake valve opens before the piston reaches TDC during the expansion stroke, fresh air is sucked into the cylinder during the intake stroke after the exhaust valve closes, and the piston enters the compression stroke after passing BDC of the intake stroke. After the intake valve is closed, compression and ignition are performed with both intake and exhaust valves closed, and the expansion stroke BD
Before C, the exhaust valve opens and the above process is repeated. When the exhaust valve opens (EO) at the beginning of the exhaust stroke, the cylinder is filled with high-pressure burnt gas, which flows out from the cylinder chamber to the exhaust port as the exhaust stroke progresses. The upward movement of the piston during the exhaust stroke facilitates this exhaustion.

However, at the end of the exhaust stroke, the intake valve opens and valve overlap occurs, so a portion of the high-pressure burnt gas flows back not only to the exhaust port but also to the low-pressure intake port. When the exhaust valve closes (EC) and the intake stroke begins, this backflow burnt gas flows into the cylinder chamber again from the intake port and mixes with fresh air in the cylinder chamber. Therefore, the amount of burned gas that flows back into the intake port, that is, the amount of burned gas that will remain in the cylinder in the next stroke, increases as the nolbuono/wrap period increases. It can be seen that the more the valves open during the high season), the more they increase.

In this example, when the load is high, that is, when cylinder combustion is mainly homogeneous mixture combustion, as shown in Fig. 5 (
The intake valve opening timing is set so that the valve overlap is relatively small as shown in a), and when stratified mixture combustion is performed at low to medium loads, as shown in Fig. 5 (b). The valve overlap is increased by advancing the opening timing of the intake valve.

By increasing the valve overlap at medium to low loads, the amount of high-temperature residual burnt gas in the cylinder increases, so the temperature of the air-fuel mixture after fresh air intake increases.

This accelerates the vaporization of the fuel and increases the combustion speed, so the problem of combustion instability mentioned above in the case of stratified mixture combustion is solved. In addition, in engines with normal throttle valves, the intake air N is throttled at low loads, so if the valve overlap is increased, the amount of fresh air will relatively decrease due to the increase in the amount of residual burnt gas, causing combustion instability. However, in an engine that does not control the amount of intake air as shown in Example 15), it is necessary to increase the amount of residual burnt gas to reduce the amount of pre-ignition gas in the cylinder chamber, since the excess air ratio is already very large. Increasing the temperature improves the combustion state.

Next, FIG. 6 shows a variable Hartz timing device 30 used in this embodiment. In this embodiment, the variable valve timing device 30 described in Japanese Unexamined Patent Publication No. 135310/1982 is used, but other types of variable valve timing devices can also be used.

JP 58-1.353 using Figures 6 to 8 below.
The variable valve control device disclosed in Japanese Patent No. 10 will be explained. FIG. 6 is a cross-sectional view showing the structure of a variable valve timing device, and in this embodiment, this variable valve timing device is used for an intake valve. In the figure, 201 is a timing gear for driving the camshaft, which is rotationally driven from the crankshaft by a belt (not shown). Also, 202 is a camshaft that drives the intake valve, and the timing gear 201 and camshaft 2.
02 are fixed to the outer sleeve 203 and inner leaf 204, respectively. The outer sleeves 203 and 204 are attached to be rotatable relative to each other, and each sleeve has a pair of slits 205.205B and 206.2 at axially symmetrical positions on the outer circumference, as shown in FIG.
06B is drilled. The slit 205 on the outer sleeve and the slit 206 on the inner sleeve are the eighth
As shown in the figure, the slits 205 and 206 are formed so as to be inclined in opposite directions with respect to the axial direction of the camshaft 202.
Roller bearings 207 and 208, which can rotate independently of each other, are provided at the intersection with each other so as to contact one of the inner walls of the sleeves h 205 and 206, respectively.

Here, roller bearings 207B and 208B are also in contact with the inner walls of the sleeves 205B and 206B (located in an axially symmetrical position to that shown in FIG. It is the opposite of the picture.

Therefore, roller bearings 207, 208 and 207B,
By simultaneously moving 208B back and forth in the axial direction of the camshaft, outer sleeve 203 and inner sleeve 204 can be rotated relative to each other without causing backlash. As mentioned above, the outer sleeve 203 is integrally fixed to the timing gear 201, and the inner sleeve 204 is integrally fixed to the camshaft 202. Therefore, the relative rotation between the outer sleeve 203 and the inner sleeve 204 causes the timing gear 201 and the camshaft 202 to rotate. As a result, the phase angle of the valve opening/closing timing with respect to the crankshaft can be changed.

Next, the roller bearings 207 (B) and 208 (B)
) along the axial direction of the camshaft will be explained. As shown in FIGS. 6 and 7, the roller bearing 207 (Bl 20g (B)
The slider 209 is rotatably supported by a bearing shaft 211 inserted into and supported by the diameter member through hole 210 of the slider 209.
is rotatably supported via a bearing 213 by a drive sleeve 212 which is movable in the axial direction but is fixed in the rotational direction. The bearing 213 has an aucle race and an inner race fixed to the slider 209 and the drive sleeve 212, respectively. Therefore, when the drive sleeve 212 moves in the axial direction without rotating, the slider 209 and the bearing shaft 211 move to the outer sleeve 203.
, can be moved in the axial direction while being rotatably held around the slider 209 together with the inner sleeve 204.

Further, a thread is provided on the inner surface of the drive sleeve, and a step motor 215 fixed to the housing 214 is connected to the drive sleeve.
It is threadedly engaged with the thread of the output shaft 216. Drive sleeve 2
12 is fixed against rotation by an axial spline 217 provided in the housing, so when the step motor rotates the output shaft 216, the drive sleeve 212 moves in the axial direction. That is, by rotating the step motor 215, the slider 209 and the o-rails 71J' 207 (B), 208 (B) are moved in the axial direction by a distance corresponding to the rotation of the step motor 215, and the outer sleeve 203 and the inner sleeve are moved. The phase of the camshaft 202 with respect to the crankshaft can be changed by relatively rotating the camshaft 204. As is clear from the above description, in this embodiment, the opening timing and closing timing of the intake valve change at the same time by the same phase, and the opening period is kept constant.

In this embodiment, as shown in Fig. 9, the intake valve opening/closing timing is advanced in a region where the engine is operated at relatively low load and low torque, that is, in a region where fuel injection is performed during the compression stroke. The Ohala lump is increased to improve combustion conditions during stratified mixture combustion.

Next, using the flowchart in FIG. 10, [E
The control operation by the CU 20 will be explained.

FIG. 10 shows a fuel injection control routine, and this routine is executed by interruption at every fixed crank angle.

In the figure, first in step 100, the engine speed N
e and the accelerator opening degree θ are read.

In step 102, from map 1 (Fig. 11), Ne
The required fuel injection amount QT is calculated based on and θ. Referring to FIG. 11, Q increases as θ increases, and Ne has a maximum value near 4000 rpm.

In step 104, it is determined whether the required fuel injection amount Q'r is equal to or less than Ql (see FIG. 4). If the determination is affirmative, the entire required fuel injection amount QT is injected in the compression stroke from step 106 onwards. That is, step 106
At step 108, the compression stroke fuel injection amount Qc is set to the required fuel injection amount Q'r, and at step 108, the intake stroke fuel injection amount Q is set to 0.
It is said that In step 110, the compression stroke fuel injection period T is started.

is Q from Map 2 (Figure 12). Calculated based on. Referring to FIG. 12, T. increases linearly as Qc increases. In step 112, the intake stroke fuel injection period T is set to O. In step 114, the compression stroke fuel injection amount Q is determined from Map 3 (Fig. 13). and the compression stroke fuel injection start timing T based on the engine speed Ne.

1 is calculated, and in step 115, the flag θADV is set to 1, and this routine ends. Referring to FIG. 13, Tel is shown as an injection advance angle from compression top dead center.

Tc1 is Ne and Q. It becomes necessary as the number increases. As will be described later, θADV is a flag representing the intake valve opening/closing timing, and when θADV-1, the intake valve opening/closing timing is advanced by a predetermined amount.

If a negative determination is made in step 104, then step 1
At step 16, it is determined whether the required fuel injection amount is less than or equal to (see FIG. 4). If affirmative determination is made, step 118
In the following, the required fuel injection it QT is divided into a compression stroke and an intake stroke. In step 118, the compression stroke fuel injection amount Q is determined. Q. 1 (see Figure 4) is entered. In step 120, the intake stroke fuel injection amount Q is changed to Q□-Q. 1
can be entered. In other words, the intake stroke fuel injection amount is 0. and compression stroke fuel injection amount Q.

The sum of the required injection amount Q is made to be the required injection amount Q. Steps 122 and 124 are map 2 (Figure 12) to Q. .

The compression stroke fuel injection period T based on Q. and the intake stroke fuel injection period T are calculated.

Next, in step 126, the compression stroke fuel injection start timing T is determined from map 4 (FIG. 14) based on the required injection amount Q and the engine speed Ne. l is calculated. Referring to FIG. 14, TCI is shown by the injection advance angle from compression top dead center, and TCI is determined by Ne and 0. As the number of people increases, it becomes necessary. In step 128, the intake stroke fuel injection start timing TSI is calculated from map 5 (FIG. 15) based on the intake stroke fuel injection amount QS and the engine speed Ne. Referring to FIG. 15, TSI is shown as an injection advance from compression top dead center, and TSI is advanced as Ne increases, but is not changed by Q5. This is because in intake stroke injection, there is sufficient time for the fuel to diffuse and form an air-fuel mixture, so there is no need to change it depending on the amount of Q. After completing the above steps, the flag θADV is set to 1 in step 129, and the routine ends.

If the determination in step 116 is negative, step 130
In the following, the entire required fuel injection amount Q1 is injected during the intake stroke. That is, in step 130, the intake stroke fuel injection amount QS is set to the required fuel injection amount QT, and in step 1
32 is the compression stroke fuel injection amount Q.

is taken to be 0. Further, in step 134, the compression stroke fuel injection period T is started. is set to O, and in step 136, the intake stroke fuel injection period TS is determined based on Q3 from map 2 (Fig. 12).
is calculated. Next, in step 138, the intake stroke fuel injection start timing TS+ is calculated from map 5 (FIG. 15) based on 0, and Ne, and in step 139, the flag θ
A[]V is reset to 0 and the routine ends.

After the above steps are completed, fuel injection is performed by another routine (not shown). Similarly, the ignition timing is set based on the required fuel injection amount Q7 and the engine speed Ne by another routine not shown.

Next, FIG. 16 is a routine showing the opening/closing timing (overlap) control of the intake valves. This routine is executed by interruption at every fixed crank angle.

In the figure, in step 150, it is determined whether θ2 and IIV are 0, and θa. If ヮ is 0, the intake valve opening/closing timing ' is not advanced and the intake valve opening timing θ1 is set to a predetermined crank angle θ8 (step 152). θ4 is an angle at which the amount of half-force exhaust of each exhaust valve becomes relatively small (see FIG. 5(a)), and when θ4 is set to 02, the amount of residual burnt gas in the cylinder chamber decreases. Further, if θADV is 1 in step 150, the intake valve opening/closing timing is advanced and the intake valve opening timing is set to a predetermined crank angle θ5 (step 154).

As a result, the intake valve opening timing is advanced by a predetermined amount, and the overlap of the intake and exhaust valves increases (see Fig. 5 (b)).
The amount of residual burnt gas in the cylinder chamber increases to stabilize the combustion of the stratified mixture.

When the setting of θ1 is completed as described above, in step 156, θ1 is set.
The value of 1 is output to the drive circuit of the variable valve timing device, and the step motor 21 of the variable valve timing device 30 is
5 to change the intake valve opening timing.

In this embodiment, the intake valve opening/closing timing is fixed in two ways, the normal value θ8 and the advance value θ8, but the intake valve opening/closing timing is changed depending on the compression stroke fuel injection amount or the intake stroke fuel injection amount. may be changed continuously.

〔Effect of the invention〕

The in-cylinder direct injection spark ignition engine according to the present invention increases the overlap of the intake and exhaust valves when the engine is under low load compared to when the engine is under high load, thereby stabilizing stratified temperature amixture combustion and stabilizing stratified temperature amixture combustion during low load operation. It has an excellent effect of increasing output and improving exhaust gas properties.

[Brief explanation of drawings]

Fig. 1 is an overall configuration diagram showing an embodiment of an engine to which the present invention is applied, Fig. 2 is a diagram showing the air-fuel mixture state in the cylinder chamber during compression stroke fuel injection, and Fig. 3 is a diagram showing the air-fuel mixture state in the cylinder chamber during split injection. Figure 4 is a diagram showing the fuel injection amount of compression stroke injection and intake stroke injection, Figure 5 is a diagram showing valve overlap, and Figures 6 to 8 are diagrams showing the variable valve timing device. Figure 9 is a diagram showing the relationship between the fuel injection pattern and valve overlap setting, Figure 10 is a flowchart showing the fuel injection control operation, and Figures 1 to 15 are the diagrams shown in Figure 10. The map used for setting control parameters, FIG. 16, is a flowchart showing valve overramp control operation. 1... Engine 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...Ignition Plug, 20...Electronic control unit, 30...Variable hull swimming device.

Claims (1)

[Claims] 1. Equipped with a fuel injection valve that can directly inject fuel into the cylinder of the engine, and when the engine load is below a predetermined value, at least a part of the required fuel injection amount according to the engine operating state is injected into the cylinder and ignited. In direct injection spark ignition engines that form a mixture near the stopper, a variable valve timing device is installed on at least one of the intake and exhaust valves, so that when the engine is running at low load, the intake and exhaust valves are set at a lower speed than when the engine is running at high load. An in-cylinder direct injection spark ignition engine, characterized in that the overlap between the two is increased.