JP2002227680A - Compression ignition type engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To reduce noise by keeping the pressure difference between the pressure inside a combustion chamber when an intake valve is opened and the pressure on an intake port side negligible all the time, and decreasing pulsation caused by blow-back. SOLUTION: At the time of the compressed ignition combustion, a negative valve overlap period when both intake and exhaust valves 6, 7 are closed together around the exhaust top dead point is formed by delaying the angle of the valve opening time of the intake valve 6, while advancing the angle of the valve closing time of the exhaust valve 7. The valve opening time of the intake valve 6 is set on the delayed angle side than the crank angle in a symmetrical position holding the valve closing time and the exhaust top dead point of the exhaust valve 7 in between. The pressure difference between the pressure inside the combustion chamber 3 when the intake valve 6 is opened and the pressure on the intake port 4 side is reduced by setting the valve opening period of the intake valve 6 on the delayed angle side than the valve closing period of the exhaust valve 7, thereby suppressing the blow-back by a residual gas.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a compression ignition engine in which a mixture in a combustion chamber is ignited at multiple points by heat of compression.

[0002]

2. Description of the Related Art Conventionally, as a part of purifying exhaust gas, for example, in a gasoline engine, various techniques for performing not only operation by spark ignition but also self-ignition operation by compression heat depending on an operation region have been studied. By using compression ignition instead of spark ignition, so-called multipoint ignition combustion with an infinite number of spark plugs is realized, so that the combustion period is short and stable combustion even at a leaner air-fuel ratio. Can be obtained. Further, by performing multipoint ignition, combustion at a low temperature becomes possible, and the emission amount of NOx can be greatly reduced.

[0003] As an engine adopting the compression ignition combustion, for example, in Japanese Patent Application Laid-Open No. 2000-320333, in the compression ignition operation region, both the intake valve and the exhaust valve are closed before and after the top dead center of the exhaust to seal the engine. A technique has been disclosed in which a negative valve overlap period is formed to increase the temperature of the air-fuel mixture with residual gas to promote compression ignition.

In this prior art, in the intake stroke, after the air-fuel mixture expands to near atmospheric pressure, the intake valve is opened to introduce fresh air.

[0005]

However, according to the technology disclosed in the above-mentioned prior art, the intake valve is opened after the pressure in the combustion chamber is expanded to near the atmospheric pressure.
The differential pressure between the pressure on the intake port side and the pressure in the combustion chamber is reduced to suppress the occurrence of "blow-back" in which residual gas in the combustion chamber is pushed back to the intake port. And it is extremely difficult to always set the pressure in the combustion chamber at the time of opening the intake valve close to the atmospheric pressure.

In view of the above circumstances, the present invention can always set the pressure difference between the pressure in the combustion chamber when the intake valve is opened and the pressure on the intake port side to be very small, attenuate the pulsation caused by the blowback and reduce the noise. And to provide a compression ignition engine capable of improving the charging efficiency.

[0007]

SUMMARY OF THE INVENTION In order to achieve the above object, the present invention provides a compression system capable of forming a negative valve overlap period for closing both an exhaust valve and an intake valve before and after a top dead center of the exhaust gas. In the ignition type engine, the opening timing of the intake valve after the exhaust top dead center is determined by the closing timing of the exhaust valve before the exhaust top dead center and the crank angle at a position symmetrical with respect to the exhaust top dead center. Is also set to the retard side.

In such a configuration, the opening timing of the intake valve when the negative valve overlap period has elapsed is retarded from the crank angle at a position symmetrical with respect to the closing timing of the exhaust valve and the exhaust top dead center. , The pressure difference between the pressure in the combustion chamber when the intake valve is opened and the pressure on the intake port side is reduced, and the blowback of the residual gas is suppressed.

In this case, preferably, 1) an in-cylinder injection means for directly injecting fuel into the combustion chamber is provided, and during low load operation, stratified charge ignition is delayed by delaying the fuel injection timing.

2) At least the intake valve is connected to a valve mechanism capable of variably setting the opening and closing timing.

[0011]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of the present invention will be described below with reference to the drawings. FIG. 1 shows an overall configuration diagram of a compression ignition type engine.

1 is an engine body, 2 is a piston, 3 is a combustion chamber, 4 is an intake port, 5 is an exhaust port, 6
Denotes an intake valve, 7 denotes an exhaust valve, and a throttle valve 9 is interposed in an intake passage 8 communicating with the intake port 4.

An injection hole of an in-cylinder injector 11 as a fuel injection means is provided at the center of the top surface of the combustion chamber 3, and a piston 2 which is opposed to the injection direction of the in-cylinder injector 11 is provided. A curved concave-shaped piston cavity 2a is formed on the top surface. Further, an ignition portion of the ignition plug 12 faces one side of the combustion chamber 3. Also, the intake valve 6
And the exhaust valve 7 are connected to the variable valve mechanisms 13a and 13b, respectively.

Each of the variable valve mechanisms 13a and 13b has
A stepped cam mechanism for switching between a plurality of cam ridges having different profiles, a continuously variable valve mechanism using an electromagnetic valve or the like is employed. Reference numeral 16 denotes a knock sensor, 17 denotes a water temperature sensor, and 18 denotes an O2 sensor.

Signals detected by these sensors are input to an electronic control unit (ECU) 20. An electronic control unit (ECU) 20 includes a CPU 21, a ROM 22, a RAM
23, an input port 24, an output port 25, etc., and are connected to each other by a bidirectional bus 26.

The input port 24 is connected to a crank angle sensor 31 for generating a crank pulse for each set crank angle in addition to the above sensors, and generates an output voltage proportional to the amount of depression of an accelerator pedal 32. A load sensor 33 is connected via an A / D converter 34. The output port 25 is individually connected to each of the variable valve mechanisms 13a and 13b via drive circuits 36a and 36b.

An electronic control unit (ECU) 20 determines an operation range based on an engine speed Ne calculated based on a signal from a crank angle sensor 31 and an engine load Lo detected by a load sensor 33, and determines the operation range. The combustion mode corresponding to the region is selected. As shown in FIG. 4, in the present embodiment, the operation region is a stratified compression ignition region I in a low-medium rotation and low load region including idling operation, a homogeneous mixed compression ignition region II in a low-medium rotation region and a medium load region, Other than high-speed or high-load region spark ignition region III,
Fuel injection timing, intake valve 6, exhaust valve 7 for each operation area
Is variably controlled.

As shown in FIGS. 3A and 5, in the stratified compression ignition region I, the valve timing is set to a negative value at which both the valves 6 and 7 close before and after the exhaust top dead center (TDC). The variable operation of the variable valve mechanisms 13a and 13b is controlled so that a valve overlap period is formed, and the fuel injection timing is set to a relatively late timing in the latter half of the compression stroke.
Further, in the homogeneously-mixed compression ignition region II, the valve timing is controlled so that a negative valve overlap period is formed before and after the exhaust top dead center (TDC), as in the stratified compression ignition region I, and the fuel injection timing is adjusted. It is set from a time after the start of the negative valve overlap period to a relatively early time during the intake stroke.

In the spark ignition region III, as shown in FIG. 3B, the valve timing is set at the top dead center (TDC) of the exhaust gas.
Before and after, the control is switched to the normal control that forms a positive valve overlap period in which the intake valve 6 and the exhaust valve 7 are both opened, and in the fuel injection control, the combustion mode (stratified combustion or It switches to the injection timing corresponding to (uniform combustion).

The control for each operation region will be described with reference to a region-specific control routine shown in FIG. In this routine,
First, in step S1, the operating range is set to the engine speed Ne.
Based on the engine load Lo and the engine load Lo, a check is made with reference to an operation region map shown in FIG.

The operating region is a stratified compression ignition region I.
If it is, the process proceeds to step S2, and if it is in the uniform mixed compression ignition region II, the process proceeds to step S3, where the spark ignition region II
If it is, the process proceeds to step S4.

When it is determined that the operation region is in the stratified compression ignition region I and the process proceeds to step S2, the variable valve mechanism 13a,
13b, a drive signal for delaying the opening timing of the intake valve 6 and advancing the closing timing of the exhaust valve 7 is output, and before and after the exhaust top dead center (TDC), the valves 6 and 7 are driven. Forms a negative valve overlap (see FIG. 3A) in which the valve closes.

That is, when the variable valve mechanisms 13a and 13b are stepped cam mechanisms, a signal for selecting a cam ridge forming a negative valve overlap is output. A valve opening / closing signal is output at the timing when the valve overlap is formed. At this time, the most effective negative valve overlap period for stratified charge compression combustion is obtained by referring to a map or the like, and the cam ridge of the stepped cam mechanism or the valve timing of the continuously variable valve mechanism is set.

If it is determined that the operating region is in the homogeneously-mixed compression ignition region II and the process proceeds to step S3, the process proceeds to step S2.
In the same manner as described above, a drive signal for forming a negative valve overlap period is output to the variable valve mechanisms 13a and 13b. At this time, a negative valve overlap period which is the most efficient for uniform mixed compression ignition combustion is obtained by referring to a map or the like, and the cam ridge of the stepped cam mechanism or the valve timing of the continuously variable valve mechanism is set.

As shown in FIG. 3A, in this embodiment, the valve opening timing IVO of the intake valve 6 when the operation region is in the stratified compression ignition region I or the homogeneously mixed compression ignition region II.
With respect to the valve closing timing EVC of the exhaust valve 7, an angle obtained by delaying the valve opening timing EVC by an angle θT from an angle SEVC symmetrical with respect to the exhaust top dead center (TDC). Is set to

The opening angle IVO of the intake valve 6 is retarded from Sevc, which is an angle symmetrical with respect to the exhaust top dead center (TDC), so that the combustion chamber 3 when the intake valve 6 is opened is opened. And the pressure difference between the internal pressure and the pressure of the intake port 4 can be reduced, and the blowback when the intake valve 6 is opened can be suppressed.

This state will be described with reference to a PV diagram schematically showing a change in pressure in the combustion chamber 3 shown in FIG.
When the exhaust valve 7 is opened during the exhaust stroke, the in-cylinder pressure changes at a pressure slightly higher than the atmospheric pressure Po due to exhaust resistance.
By closing the exhaust valve 7 in the middle of the exhaust stroke (since the intake valve 6 is also closed at this time), the residual gas having a pressure slightly higher than the atmospheric pressure Po is trapped in the combustion chamber 3. As the piston 2 rises, the preload temperature is raised.

Then, the piston 2 is moved to the exhaust top dead center (TD).
C), the process proceeds to the intake stroke, and when the intake valve 6 is opened at the valve opening timing IVO, fresh air is introduced into the combustion chamber 3. This fresh air is heated and heated by the residual gas, and is also heated by the subsequent compression stroke.

Since the valve opening timing IVO of the intake valve 6 is delayed from the valve closing timing EVC of the exhaust valve 7 by an angle θT with respect to an angle Sevc symmetrical with respect to the exhaust top dead center TDC, the intake valve The in-cylinder pressure P1 when the valve 6 is opened is lower than the in-cylinder pressure P2 when the exhaust valve 7 is closed. Therefore, the differential pressure ΔP1 between the atmospheric pressure Po and the in-cylinder pressure P1 is determined by the exhaust valve 7 It shows a value smaller than the differential pressure ΔP2 when the valve is closed. on the other hand,
Since the intake port 4 indicates a pressure almost equal to the atmospheric pressure Po, even if the intake valve 6 is opened, residual gas is not removed from the intake port 4.
"Blowback" pushed back to the side does not occur, and noise caused by the blowback and reduction in filling efficiency can be effectively avoided.

On the other hand, for example, when the intake valve 6 is opened at an angle Sevc symmetrical with respect to the closing timing EVC of the exhaust valve 7 with respect to the exhaust top dead center (TDC), the in-cylinder pressure P2
Is higher than the atmospheric pressure Po, the blowback occurs due to the pressure difference ΔP2.

If the retard amount θT is too large, the in-cylinder pressure P
Is lower than the reference pressure P0, and a pumping loss occurs. Therefore, the optimal retard amount θT is obtained from an experiment or the like, and the opening timing of the intake valve 6 is set based on the retard amount θT.
In this case, when a continuous variable valve mechanism is employed as the variable valve mechanisms 13a and 13b, the retard amount θT may be variably set according to the operating state.

When the process proceeds from step S2 to step S5, a fuel injection timing at which stratified charge ignition combustion can be performed is set, and the routine exits. Also, from step S3 to step S
In step 6, the fuel injection timing at which homogeneous charge compression ignition combustion is possible is set, and the routine exits.

As shown in FIG. 5, the fuel injection timing during stratified charge compression ignition combustion is set to a relatively late time in the latter half of the compression stroke. Further, the fuel injection timing at the time of the homogeneous charge compression ignition combustion can be set as appropriate from the time when the exhaust valve 7 is closed (the negative valve overlap period starts) to the end of the intake stroke.

In the stratified compression ignition region I, by injecting fuel from the in-cylinder injector 11 during the compression stroke, a stratified fuel with a high fuel concentration enters the combustion chamber 3 which is reaching a gas temperature at which compression ignition combustion is possible. Since the liquefied mixture is locally generated, compression ignition combustion at an extremely lean air-fuel ratio becomes possible.

In the homogeneous charge compression ignition region II, the in-cylinder injector 11 is operated relatively early after the exhaust valve 7 is closed.
By injecting fuel from the combustion chamber, a uniform mixture can be generated before the gas temperature in the combustion chamber 3 reaches the self-ignition possible temperature, and when the self-ignition temperature is reached, the mixture in the combustion chamber 3 is simultaneously blown. Ignited and the flame does not propagate, that is, multipoint ignition combustion in which an infinite number of spark plugs are arranged.

In this case, the combustion temperature of the air-fuel mixture by the self-ignition combustion is about 1800 ° C., which is lower by about 200 ° C. than that of the combustion by the normal ignition. Become. As a result, the amount of NOx emission is reduced, and the ignition is performed all at once. Thus, the thermal efficiency is good, and the exhaust gas can be made essentially clean while maintaining the fuel efficiency of a direct injection diesel engine.

Incidentally, gasoline or methanol having a low cetane number is employed as the fuel. Even with such a low cetane number fuel, for example, in the case of gasoline, it is known that when the gas temperature in the combustion chamber 3 exceeds about 900 ° C., the fuel self-ignites. The temperature rise at the time of pre-pressurizing the gas and the normal compression ratio are used as temperature rise means, and the function is set so that the gas temperature reaches the self-ignition possible temperature by these functions.

That is, the theoretical thermal efficiency ηth is as shown in the following equation in the case of a constant volume cycle. ηth = 1− (1 / εκ−1) Here, ε: compression ratio, κ: specific heat ratio.

Therefore, the gas temperature can be raised to a self-ignition possible temperature within a practical range without significantly increasing the compression ratio ε by increasing the preload temperature of the residual gas. As a result, even if the compression ratio ε is set to ε = 12 to 15, sufficient compression ignition combustion can be obtained, and even if the control shifts to spark ignition control, torque reduction due to the knocking limit occurs. It becomes difficult.

On the other hand, if it is determined in step S1 that the ignition region is the spark ignition region III and the process proceeds to step S4, the variable valve mechanism 13
In response to a and 13b, a drive signal is output in which the valve timing is a positive valve overlap period (see FIG. 3B) in which the intake valve 6 and the exhaust valve 7 are both opened from the end of the exhaust stroke to the beginning of the intake stroke. I do. That is, the variable valve mechanism 13
When a and 13b are stepped cam mechanisms, a signal for selecting a cam ridge forming a positive valve overlap is output.
In the case of a continuously variable valve operating mechanism, a valve opening / closing signal at a timing for forming a positive valve overlap period is output.

Next, the routine proceeds to step S7, in which spark ignition control is executed, and the routine is exited. Since the fuel injection control and the ignition timing control at the time of the spark ignition control are known, their description is omitted here. In this case, the operation region may be subdivided into a stratified combustion region and a uniform combustion region, and the fuel injection timing may be variably set for each region.

As described above, according to the present embodiment, the top dead center (TDC) of the exhaust valve 7 is set with respect to the valve opening timing IVO of the intake valve 6 and the valve closing timing EVC of the exhaust valve 7 during the compression ignition combustion. Since the retardation is retarded by the retard amount θT from the symmetrical angle Sevc, the differential pressure between the pressure in the combustion chamber 3 and the pressure on the intake port 4 side when the intake valve 6 opens is reduced. Blowback can be effectively prevented.

Further, since the opening timing of the intake valve is set based on the closing timing of the exhaust valve, the differential pressure when the intake valve is opened can always be made small.

[0044]

As described above, according to the present invention,
The opening timing of the intake valve during compression ignition combustion is set to be more retarded than the crank angle at a position symmetrical with respect to the closing timing of the exhaust valve and the exhaust top dead center. The differential pressure between the pressure in the combustion chamber and the pressure on the intake port side is reduced, the pulsation due to the blowback is attenuated, the noise is reduced, and the charging efficiency is improved.

Further, since the opening timing of the intake valve is set based on the closing timing of the exhaust valve, the differential pressure when the intake valve is opened can always be made small.

[Brief description of the drawings]

FIG. 1 is an overall configuration diagram of a compression ignition engine.

FIG. 2 is a flowchart showing an area-specific control routine;

3A is an explanatory diagram showing valve timing during compression ignition combustion, and FIG. 3B is an explanatory diagram showing valve timing during spark ignition.

FIG. 4 is an explanatory diagram of an operation area map.

FIG. 5 is an explanatory diagram showing a relationship between valve timing and fuel injection timing.

FIG. 6 is a PV diagram showing a change in pressure in the combustion chamber.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Engine main body 3 Combustion chamber 6 Intake valve 7 Exhaust valve 13a Variable valve mechanism EVC Closed angle of exhaust valve IVO Open angle of intake valve θT Delay angle TDC Exhaust top dead center

 ──────────────────────────────────────────────────続 き Continuing on the front page (72) Inventor Hitoshi Ito 3-9-6 Osawa, Mitaka City, Tokyo Inside Subaru Research Institute Co., Ltd. (72) Inventor Yohei 3-9-6 Osawa, Mitaka City, Tokyo F-term in Subaru Research Laboratories (reference) HA04 HA16 HA19 JA25 JA37 KA07 KA08 KA09 KA24 KA25 LA07 LB04 MA19 NC02 NE12 PE01Z PE03Z PF03Z

Claims (3)

[Claims] 1. A compression ignition engine capable of forming a negative valve overlap period in which both an exhaust valve and an intake valve are closed before and after an exhaust top dead center, wherein the intake air after the exhaust top dead center is provided. Compression ignition, wherein a valve opening timing of the valve is set to a more retarded side than a crank angle at a position symmetrical with respect to a closing timing of the exhaust valve before the exhaust top dead center and the exhaust top dead center. Expression engine. 2. A compression ignition type engine according to claim 1, further comprising in-cylinder injection means for directly injecting fuel into the combustion chamber, and stratifying compression ignition by delaying fuel injection timing during low load operation. 3. A compression ignition type engine according to claim 1, wherein at least the intake valve is connected to a valve operating mechanism capable of variably setting an opening / closing timing.