JP4529764B2 - 4 cycle spark ignition engine control device for vehicle - Google Patents

Description

  The present invention relates to a control device for a four-cycle spark ignition engine for a vehicle.

  In order to improve the output of the engine, it is effective to increase the compression ratio.

  However, simply increasing the geometric compression ratio of the engine causes a problem that knocking is likely to occur.

In order to solve such a problem, for example, in the technique described in Patent Document 1, a technique is disclosed in which fuel is dividedly injected and a subsequent fuel injection is performed after ignition of the preceding fuel.
Japanese Patent Laid-Open No. 9-1226028

  In the technique of Patent Document 1, since the air-fuel ratio becomes lean by burning the separately injected fuel, knocking can be relatively suppressed.

  However, in the technique of Patent Document 1, only the method of dividing the fuel injection timing and igniting the subsequent fuel after the combustion of the preceding fuel is used, so that the torque is significantly reduced when the subsequent fuel ignition is delayed. On the other hand, if it is attempted to accelerate the subsequent fuel ignition, it becomes difficult to take measures against knocking. Thus, since it is difficult to achieve both knocking and torque, there is a limit to the compression ratio at which knocking can be avoided, and it has been difficult to improve the output in the low rotation operation region under high load.

  The present invention has been made in view of the above problems, and is a four-cycle spark ignition type for a vehicle that can suppress knocking while adopting as high a geometric compression ratio as possible and can also improve fuel efficiency. It is an object to provide an engine control device.

  In order to solve the above-mentioned problems, the present invention cooperates with the piston by fitting a crankshaft that outputs torque, a piston coupled to the crankshaft, and a reciprocating movement of the piston. In a control apparatus for a four-cycle spark-ignition engine for a four-cycle spark-ignition vehicle having a cylinder that defines a combustion chamber for an air-fuel mixture, the geometric compression ratio of the engine is set to 14 or more, and the engine is operated. An operating state detecting means for detecting the state, a plurality of spark plugs per cylinder, and a control means for controlling the operation of the spark plug based on the detection of the operating state detecting means are provided, and the engine speed is higher than the idle speed by a predetermined number of revolutions. In the high load operation region in the engine low rotation region, the control means controls the spark plug so that multipoint ignition is performed after the compression top dead center. Dual as the control device for the four-cycle spark ignition engine. In this aspect, it is possible to obtain high output performance by setting the geometric compression ratio of the engine to 14 or more. In addition, by setting the ignition timing in the high load operation region in the low engine speed region, which is higher than the idling engine speed by a predetermined number of revolutions, after the compression top dead center, the knocking limit is shifted, thereby preventing knocking. It becomes possible. By the way, when the ignition timing is set after the compression top dead center, the combustion usually occurs in the expansion stroke in which the piston descends, so the combustion speed becomes too slow and the fuel energy is effectively converted into expansion work. The output performance is greatly reduced. On the other hand, in the present invention, it becomes possible to increase the combustion speed by using multipoint ignition, and at the high load, while maintaining the output performance equivalent to the conventional engine of the normal compression ratio, the partial compression depends on the high compression ratio. Fuel economy can be improved.

  In a preferred embodiment, intake air heating means for heating the intake air is provided, and in the partial load operation region, the control means controls the intake air heating means so as to heat the in-cylinder intake air and perform the compression self-ignition operation. In this aspect, in the operation region where knocking is relatively difficult to occur, compression self-ignition can be executed to further improve fuel efficiency.

  In a preferred embodiment, on the high load side of the partial load operation region, the control means controls the spark plug so that multipoint ignition is performed after compression self-ignition. When compression self-ignition is executed at a low load, it is a problem to smoothly switch to normal spark ignition on the high load side. In order to increase the load, it is necessary to control the increase of fresh air while suppressing the temperature rise due to intake air heating and internal EGR, but this also makes it difficult for compression self-ignition to occur, and misfire is mixed. This leads to a stable area. Therefore, in this mode, multipoint ignition is executed at a timing later than the timing at which compression self-ignition occurs (after top dead center), and combustion is established by ignition even in a cycle where compression self-ignition has not been established. Misfire and deterioration of emission are prevented in the vicinity of the relatively high load operation region where the self-ignition operation is switched to the spark ignition operation.

  In another aspect, an external EGR device that recirculates a part of burned gas to the intake passage is provided, and external EGR is introduced in the partial load operation region and compression top dead center in a relatively high load operation region of the EGR introduction region. The control means controls the external EGR device and the spark plug so that multipoint ignition is performed later. In other words, on the low load side, multipoint ignition may be executed before the compression top dead center. In this aspect, since the geometric compression ratio is high, knocking is likely to occur from the partial load operation region. However, the introduction of external EGR increases the specific heat and suppresses the temperature rise of the air-fuel mixture, thereby suppressing knocking. . Furthermore, the rapid combustion effect of multi-point ignition is added, the knocking suppression effect is increased, the knocking suppression region can be expanded at the time of partial load, and the fuel efficiency can be improved at the same time.

  In a preferred embodiment, the spark plug is set to at least three or more including those laid out in the center of the cylinder. In this aspect, when performing multi-point ignition, one spark plug ignites the air-fuel mixture from the center portion of the cylinder, and the remaining spark plug ignites the air-fuel mixture from other parts. In addition to being quick, even when multipoint ignition is not performed, the air-fuel mixture can be combusted relatively quickly by adopting the spark plug disposed in the center portion.

  In a preferred aspect, in the engine, the cylinder bore center of the cylinder is offset to the right side from the rotation center of the crankshaft when viewed from the side in which the rotation direction of the crankshaft is clockwise. In this aspect, when the cylinder bore center of the cylinder is offset from the rotation center of the crankshaft, the piston lifting speed is asymmetric with respect to the top dead center, and the piston lowering speed at the initial stage of the expansion stroke is relatively slow. . For this reason, since the combustion speed becomes relatively faster than the descending speed of the piston, it is possible to maintain a good combustion environment, and it becomes possible to increase the energy acting on the piston and improve the fuel consumption. .

  In a preferred embodiment, swirl generating means for generating a swirl is provided in the cylinder. In this aspect, the generation of swirl can effectively maintain the turbulent energy up to the top dead center, and contribute to the improvement of the combustion speed during the expansion stroke in addition to the multipoint ignition. For this reason, even when performing multipoint ignition after top dead center, efficiency deterioration can be suppressed. Also, in the compression self-ignition operation region at the time of partial load, it is effective for mixing fuel and air, and in the spark ignition operation region, it is effective for suppressing knocking due to rapid combustion.

  In a preferred embodiment, the first timing when the intake valve closing timing is a predetermined timing before bottom dead center and the effective compression ratio is substantially equal to the expansion ratio, and the second timing when the effective compression ratio is smaller than the expansion ratio. A variable valve timing mechanism that can be changed to the timing is provided, and the control means controls the variable valve timing mechanism so that the closing timing of the intake valve becomes the second timing at least when the engine is warm. In this mode, in the operation region where high engine output is required, the intake valve can be driven so that the effective compression ratio is substantially equal to the expansion ratio, and the output can be improved by the high compression ratio. At the time of warm start that requires only a low value, it is possible to improve the startability by driving the intake valve so that the effective compression ratio is low and reducing the compression resistance.

  As described above, according to the present invention, there is a remarkable effect that knocking can be suppressed and fuel consumption can be improved while adopting as high a geometric compression ratio as possible.

  Hereinafter, preferred embodiments of the present invention will be described with reference to the accompanying drawings.

  FIG. 1 is a configuration diagram illustrating a schematic configuration of a control device 10 according to an embodiment of the present invention, and FIG. 2 is a schematic cross-sectional view illustrating a structure of one cylinder of the four-cycle gasoline engine 20 according to FIG. .

  With reference to FIGS. 1 and 2, the illustrated control apparatus 10 includes a four-cycle gasoline engine 20 and a control unit 100 as a control means for controlling the engine 20.

  The engine 20 integrally includes a cylinder block 22 that rotatably supports the crankshaft 21 and a cylinder head 23 disposed on the upper portion of the cylinder block 22, and the cylinder block 22 and the cylinder head 23 include A plurality of cylinders 24 are provided.

  Each cylinder 24 is provided with a piston 26 connected to the crankshaft 21 via a connecting rod 25 and a combustion chamber 27 formed in the cylinder 24 by the piston 26. In the present embodiment, the geometric compression ratio of each cylinder 24 is set to 14.

  Referring to FIG. 2, in the engine 20 according to the present embodiment, the cylinder bore center Z of the cylinder 24 is the position of the crankshaft 21 as viewed from the side where the rotation direction of the crankshaft 21 is clockwise (that is, the state of FIG. It is offset from the center of rotation O to the right. This offset amount S is set to, for example, 1 mm to 2 mm when the bore diameter of the cylinder 24 is 70 mm.

  FIG. 3 is a model diagram of the piston in the present embodiment.

With reference to the figure, the connecting rod 25 and the piston 26 and A ₁ the distance of the rotation center of the rotation center of the connecting pin 25b that connects the connecting rod 25 and the crankshaft 21 of the connecting pin 25a for connecting the _rotation of the connecting pin 25b The distance between the center and the rotation center O of the crankshaft 21 is A ₂ , the offset amount is S, the angle between the vertical line LN1 passing through the connecting pin 25b and the connecting rod 25 is α, and the angular velocity of the crankshaft 21 is ω.

According to FIG. 3, the coordinates (Xp, Yp) of the rotation center of the connecting pin 25a and the coordinates (Xc, Yc) of the rotation center of the connecting pin 25b are the horizontal line passing through the center O of the crankshaft 21 and the cylinder bore center Z. If the intersection coordinates are (0, 0), (Xp, Yp) = (0, A ₂ cosωt + A ₁ cosα) (Equation 1)
(Xc, Yc) = (A ₂ sin ωt−S, A ₂ cos ωt) (Expression 2)
It becomes. The constraint condition of the piston 26 is
A ₂ sin ωt−S = A ₁ sin α (Expression 3)
It becomes. Here, if both sides of (Equation 3) are differentiated,

・・・（式４）
が得られる。

... (Formula 4)
Is obtained.

  Next, the Yp component of (Equation 1) is differentiated by t, and (Equation 4) is substituted to obtain the velocity v of the piston 26.

・・・（式５）
が得られる。また、（式３）より

... (Formula 5)
Is obtained. From (Equation 3)

・・・（式６）
であるから、（式６）を（式５）に代入すると、

... (Formula 6)
Therefore, if (Equation 6) is substituted into (Equation 5),

・・・（式７）
が得られる。

... (Formula 7)
Is obtained.

  FIG. 4 is a graph showing the piston speed with respect to the crank angle. FIG. 5 is a model diagram showing the relationship between the crank angle and the piston movement amount.

  4 and 5, as is clear from (Equation 5), when offset amount S = 0, since tan α = 0, the speed v of the piston 26 becomes equal to a sine wave.

On the other hand, when the offset amount S> 0, the waveform of the piston 26 becomes asymmetric according to the value of the second term of (Equation 5) and (Equation 7), and the point P ₁ where the piston 26 moves at the highest speed, P ₂ (contact point formed by a straight line passing through the top dead center of the piston 26 on the locus 25L of the connecting pin 25b) is asymmetrical. As a result, if the movement amounts Y ₁ and Y ₂ of the same amount of crank angle ωt (= 30 °) around the point P ₀ when the piston 26 is at the top dead center are considered, it is clear from FIG. Thus, the movement amount Y ₁ from the vicinity of the top dead center to the top dead center becomes relatively large, and the piston 26 moves at a high speed, whereas the movement amount Y ₂ after exceeding the top dead center. Will be relatively small and the piston 26 will move at a relatively slow speed.

  FIG. 6 is a schematic plan view showing the cylinder 24 in an enlarged manner. FIG. 7 is an explanatory view showing the air flow in the combustion chamber according to this embodiment, where (A) shows the initial stage of the compression stroke and (B) shows the initial stage of the expansion stroke.

  Referring to FIGS. 6 and 7, each cylinder 24 is formed with a ceiling portion of a combustion chamber 27 on the lower surface of the cylinder head 23, and this ceiling portion extends from the central portion to the lower end of the cylinder head 23. It is a so-called pent roof type having inclined surfaces 27a and 27b.

  Two independent intake ports 28 are opened in one inclined surface (the right inclined surface in FIGS. 7A and 7B) 27a constituting the ceiling portion of the combustion chamber 27, and the other inclined surface. (An inclined surface on the left side in FIGS. 7A and 7B) 27b has two exhaust ports 29, and an intake valve 30 and an exhaust valve 31 are provided at the open ends of the ports 28 and 29, respectively. . Each intake port 28 is a straight port extending linearly from the combustion chamber 27 to the upper right in FIG. 7, and is configured to be away from the cylinder bore center Z toward the intake upstream side in the cross section shown in FIG.

  A fuel injection valve 32 that receives a fuel injection pulse from the following control unit and injects fuel corresponding to the pulse width into the combustion chamber 27 is provided at the side of the combustion chamber 27.

  On the crown surface of the piston 26, squish area constituting surfaces 26a and 26b are provided so as to be inclined along the inclined surface of the cylinder head 23 in a predetermined range of the peripheral portion on the intake side and a predetermined range of the peripheral portion on the exhaust side. ing. Further, a recessed portion 33 is provided inside the squish area constituting surfaces 26a and 26b.

  The recessed portion 33 is provided in a predetermined range including the projection surfaces of the intake valve 30 and the exhaust valve 31. The bottom surface of the recessed portion 33 has a pair of inclined bottom surfaces 33a and 33b substantially parallel to the both inclined surfaces 27a and 27b of the combustion chamber ceiling portion. , 33b is formed in a mountain shape where the ridge line portion 33c is located. The ridge line portion 33c extends linearly in the same direction as the ridge line between the inclined surfaces 27a and 27b of the combustion chamber ceiling portion at a position slightly offset to the exhaust side from the cylinder bore center Z.

  Both the intake port side periphery (boundary with the intake side squish area constituting surface 26a) 33d and the exhaust port side periphery (boundary with the exhaust side squish area constituting surface 26b) 33e of the recessed portion 33 are parallel to the ridge line portion 33c. It is formed in a straight line. The vertical walls 33f extending from the squish area constituting surfaces 26a and 26b to the bottom surface of the recessed portion 33 at the intake port side peripheral edge 33d and the exhaust port side peripheral edge 33e are inclined in the lift direction of the intake valve 30 and the exhaust valve 31. Is formed.

  For this reason, in this embodiment, there exist the following effects in the partial load operation area | region D (refer FIG. 8).

  That is, the intake air sucked into the combustion chamber 27 by the lowering of the piston 26 during the intake stroke flows into the combustion chamber 27 mainly from the side near the ignition plug 34 of the opening of the intake port 28 toward the combustion chamber peripheral wall surface on the exhaust side. The flow then continues downward along the exhaust-side peripheral wall surface, then flows along the piston crown surface toward the intake side, and then upwards. In this way, a positive tumble flow T1 swirling counterclockwise in FIG. 7 is generated.

  Next, when moving to the compression stroke, as the piston rises, the normal tumble flow T1 is swung in the combustion chamber 27 while being compressed in the vertical direction, and the flow on the lower side enters the recess 33, and the recess 33 Move along the bottom. In this case, since the bottom surface of the recessed portion 33 is formed in a mountain shape having the ridge line portion 33c at the substantially central portion, the airflow flowing along the inclined bottom surface 33b on the exhaust side is separated from the bottom surface at the ridge line portion 33c. The Then, as shown in FIG. 7A, by separating at the ridge line portion 33c, a part of the normal tumble flow T1 is diverted, and the reverse tumble flow T2 swirling in the opposite direction to the normal tumble flow T1. Produce.

  In the initial to middle stage of the compression stroke, the normal tumble flow T1 is larger and stronger than the divided reverse tumble flow T2, but as the compression stroke progresses and the piston 26 approaches the top of the combustion chamber, the normal tumble flow T1. Becomes smaller as the center gradually moves to the exhaust side. Then, at the end of the compression stroke or the early stage of the expansion stroke where the piston 26 is in the vicinity of the top dead center, as shown in FIG. 7B, the tumble flows T1 and T2 have the same magnitude and strength, and the combustion chamber 27 The exhaust side and the intake side are separated and turn in opposite directions.

  Further, since the squish area constituting surfaces 26a and 26b are provided in a predetermined range of the intake port side peripheral portion and a predetermined range of the exhaust port side peripheral portion of the crown surface of the piston 26, the piston 26 is compressed close to top dead center. At the end of the stroke, the direction from the squish area between the inclined surfaces 27a and 27b of the ceiling portion of the combustion chamber and the squish area constituting surfaces 26a and 26b toward the center of the combustion chamber 27 (the white arrow in FIG. 7B) In the initial stage of the expansion stroke in which the forward squish flow occurs in the direction opposite to Ra and Rb and the piston 26 starts to descend after reaching the top dead center, it is indicated by white arrows Ra and Rb in FIG. Such a reverse squish flow from the center of the combustion chamber 27 toward the squish area is generated.

  In this case, since the tumble flows T1 and T2 separated into two are in the reverse direction to the normal squish flow and the same direction as the reverse squish flow Ra and Rb, the normal squish flow at the end of the compression stroke is weakened. The effect of accelerating the generation of the reverse squish flow and strengthening the reverse squish flow Ra, Rb is exhibited.

  By strengthening the reverse squish flows Ra and Rb in this way, in the partial load operation region D (see FIG. 8), the combustion speed in the squish area is sufficiently increased, and the main combustion speed of the flame is increased. Rapid combustion is realized. In addition, since the normal tumble flow T1 is moderately weakened, the initial combustion speed is not so high, and self-ignition of the air-fuel mixture in the end gas zone is not induced. That is, the initial combustion period is not so short, and the main combustion period is significantly shortened, so that the occurrence of knocking is suppressed and thermal efficiency is improved by rapid combustion.

  Further, since the pair of inclined bottom surfaces 33a and 33b are parallel to the inclined surfaces 27a and 27b of the combustion chamber ceiling portion, flame propagation is performed uniformly in the space between them, which is effective in preventing detonation.

  2 and 6, each cylinder 24 is provided with three spark plugs 34 that are fixed to the cylinder head 23 and emit a spark in the combustion chamber 27. Each spark plug 34 is arranged along the cylinder diameter parallel to the ridge line portion 33 c of the piston 26, the center one is disposed on the cylinder bore center Z, and both sides are disposed on the side edge of the combustion chamber 27. . Each ignition plug 34 is connected to an ignition circuit 35 capable of controlling the ignition timing by electronic control. By controlling the ignition circuit 35 by the control unit 100, the ignition plug 34 is selectively ignited. To be controlled.

  Next, referring to FIG. 2, each of the intake valve 30 and the exhaust valve 31 of each cylinder 24 is provided with a known tappet unit 36. The tappet unit 36 is periodically driven by cams 37 a and 38 a of camshafts 37 and 38 of a valve mechanism provided in the cylinder head 23. The valve operating mechanism for the intake valve 30 is provided with a variable valve timing mechanism 40 that can change the opening / closing timing of the intake valve 30. The variable valve timing mechanism 40 has a valve according to the operating state over a first timing in which the intake valve opening timing is near the intake top dead center and a second timing in which the intake valve opening timing is advanced. The timing is changed.

  Next, referring to FIGS. 1 and 2, a branch intake pipe 43 of the intake manifold 42 is connected to the intake port 28 of the engine 20. The branch intake pipe 43 is provided for each cylinder 24 and is connected to the intake manifold 42 in a state where an equal-length intake path is formed. In the illustrated embodiment, each cylinder 24 is formed with a pair of intake ports 28, and the downstream end of the branched intake pipe 43 is bifurcated corresponding to the intake port 28 of each cylinder 24. Has been. An open / close valve 44 is provided at the upstream side merge portion of the branch intake pipe 43. The on-off valve 44 is configured so that the collective portion of the branch intake pipe 43 can be opened and closed individually by the actuator 45. On the other hand, a known swirl on / off valve 43a is provided at one branch portion of the branch intake pipe 43 branched into two branches as shown in FIG. The swirl on / off valve 43a is driven by an actuator 43b to open and close. When one of the branch portions of the branch intake pipe 43 is closed by the swirl on / off valve 43a, the swirl on / off valve 43a passes through the other branch. The swirl is generated in the combustion chamber 27 by the intake air, and the swirl is weakened as the swirl generation opening / closing valve 43a is opened.

  An intake passage 46 for introducing fresh air into the intake manifold 42 is connected to the upstream side of the intake manifold 42. A throttle valve 47 is provided in the intake passage 46. A three-way solenoid valve 48 is provided on the upstream side of the throttle valve 47, and a heater 50 as intake air heating means is provided in a bypass passage 49 connected to the three-way solenoid valve 48. Therefore, by switching the three-way solenoid valve 48, fresh fresh air can be introduced as it is into the intake manifold 42, or air heated by the heater 50 can be introduced into the intake manifold 42. .

  Next, as shown in FIG. 1, the exhaust port 29 is connected to a bifurcated branch exhaust pipe 51 formed in pairs for each cylinder 24. The downstream end of each branch exhaust pipe 51 is connected to the exhaust manifold 52. An exhaust passage 53 for discharging burned gas is connected to the exhaust manifold 52.

  Next, referring to FIGS. 1 and 2, an external EGR system 60 is provided between the intake manifold 42 and the exhaust manifold 52 to recirculate the exhausted burned gas to the intake manifold 42.

  The external EGR system 60 is connected to a recirculation passage 61 formed between the intake manifold 42 and the exhaust manifold 52, and includes an EGR cooler 62, an EGR valve 63, and an actuator 64 that drives the EGR valve 63. Valve system.

  Referring to FIG. 1, in order to detect the operating state of engine 20, air flow sensor SW1 is provided in intake passage 46, and intake temperature sensor SW2 for predicting the in-cylinder temperature downstream of on-off valve 44. (See FIG. 2). The cylinder block 22 is provided with a crank angle sensor SW3 for detecting the rotation speed of the crankshaft 21 and an engine water temperature sensor SW4 for detecting the temperature of the cooling water (see FIG. 2). Further, the exhaust passage 53 is provided with an oxygen concentration sensor SW5 for controlling the air-fuel ratio.

  The control unit 100 of the engine 20 has an airflow sensor SW1, an intake air temperature sensor SW2, a crank angle sensor SW3, an engine water temperature sensor SW4, an oxygen concentration sensor SW5, and an accelerator opening sensor SW6 for detecting engine load as input elements. It is connected. Each of these sensors SW1 to SW6 is a specific example of the driving state detection sensor in the present embodiment. On the other hand, the control unit 100 includes an actuator of a throttle valve 47, a variable valve timing mechanism 40, an actuator 43b of a swirl generating on-off valve 43a, a three-way electromagnetic valve 48 of an intake passage 46, a heater 50, and an actuator 64 of an external EGR system 60. Connected as a control element.

  Referring to FIG. 1, a control unit 100 has a CPU 101, a memory 102, an interface 103, and a bus 104 for connecting these units 101 to 103. The control unit 100 operates according to programs and data stored in the memory 102. The operation state determination means for determining is functionally configured.

  FIG. 8 is a characteristic diagram illustrating an example of operation region setting for performing control according to the operation state according to the embodiment of FIG. 1. FIG. 9 is a timing chart showing intake valve opening timing and ignition timing for each characteristic.

  Referring to FIG. 8, the control unit 100 has a control map based on a characteristic diagram as shown in FIG.

In the characteristic diagram shown in FIG. 8, in the high load operation region D ₀ in a low rotation region (region of engine rotation speed N2 or less) N1 or higher at a predetermined rotation speed (1200 rpm in the illustrated example), FIG. As indicated by SP, the multipoint ignition is set after the compression top dead center. As a result, when the ignition timing is after compression top dead center, the knocking limit is extended to the high speed side, as will be described in detail later, and the combustion speed is accelerated by using multipoint ignition to make the fuel energy effective. It is possible to secure the output performance by converting into expansion work. As a result, while maintaining a high geometric compression ratio on the high load side in the relatively low rotation operation region where compression self-ignition operation is difficult, while preventing ignition by delaying the ignition timing, It becomes possible to improve fuel consumption.

Then, in the characteristic diagram shown in FIG. 8, the partial load operation region D (both parts of the region D ₂ indicated by the outline and the area D ₁ indicated by hatching than this low load side), in FIG. 9 (B) As shown, the heater 50 is activated and a compression self-ignition operation is performed.

Then, in the characteristic diagram shown in FIG. 8, one of the partial load operation region D, the region D ₁ of the high-load side performs ignition assistance in the same timing as FIG. 9 (A), the ignition of the compressed self-ignition It is set to assist. Further, in the operation region D ₁ on the high load side, the swirl generating on-off valve 43a is controlled by the control unit 100. As a result, swirl occurs during intake, turbulent energy is maintained up to top dead center, and mixing of fresh air introduced from the intake port 28 and burned gas introduced from the exhaust valve 31 is promoted. . For this reason, even if the ignition timing is delayed in the first half of the expansion stroke, the air-fuel mixture can be burned quickly and the deterioration of efficiency can be suppressed. Further, since the swirl is generated in the operating region D ₁ of the particularly high load side, since the mixing of the fuel is also promoted, it can also contribute to a reduction of cylinder temperature, be further avoid knocking In addition, the fuel efficiency is improved and the lean combustion limit is improved.

  Next, referring to FIGS. 9A to 9C, in each operation region described above, as shown in FIGS. 9A and 9B, the intake valve 30 is approximately at the top dead center of the intake stroke. The valve is opened at the first timing that starts and ends at the bottom dead center. On the other hand, when the engine 20 is restarted when it is warm, the variable valve timing mechanism 40 causes a considerable delay, which is shown in FIG. As shown, the effective compression ratio opens at a second timing that becomes smaller than the expansion ratio. As a result, the piston 26 that has shifted to the compression stroke performs a so-called Miller cycle function in which the internal air is blown back to the intake port 28 side.

  As described above, in the present embodiment, high output performance can be obtained by setting the geometric compression ratio of the engine 20 to 14 or more.

  FIG. 10 is a characteristic diagram showing the relationship between engine load and engine speed.

  As shown in the figure, an engine having a relatively large geometric compression ratio ε has a larger torque as shown by a curve ε1, while an engine having a relatively small geometric compression ratio ε (curve ε2). The region that reaches the knocking limit (NK1) extends to the high rotation side.

On the other hand, in the present embodiment, the ignition timing in the high load operation region D ₀ in the engine low rotation region that is higher by the predetermined rotation speed N1 than the idle rotation speed is after the compression top dead center as shown in FIG. Since it is set, the knocking limit shifts to the low speed side as indicated by the phantom line NK1 to the solid line NK2 in FIG. As a result, knocking can be reliably prevented even in the low rotation operation region. By the way, when the ignition timing is set after the compression top dead center, the combustion is normally performed in the expansion stroke in which the piston 26 descends. Therefore, the combustion speed becomes too slow and the fuel energy is effectively converted into expansion work. It becomes impossible to do so, and the output performance is greatly reduced. On the other hand, in this embodiment, as shown in FIG. 9A, the combustion speed is increased by performing multipoint ignition at the time of ignition. Therefore, the high expansion effect by the high compression ratio engine 20 can be exhibited while avoiding knocking, and the fuel consumption can be improved.

  In the present embodiment, a heater 50 is provided as intake air heating means for heating the intake air. In the partial load operation region D shown in FIG. 8, the control unit 100 is configured to heat the in-cylinder intake air and perform the compression self-ignition operation. Controls the heater 50. For this reason, in the present embodiment, it is possible to perform compression self-ignition in an operation region where knocking is relatively difficult to occur, thereby further improving fuel consumption.

Further, in this embodiment, in the operating region D ₁ of the high-load side of the partial load operation region D, the control unit 100 so as to multi-point ignition after compression self-ignition to control the ignition plug 34. For this reason, in this embodiment, multipoint ignition is executed at a timing (after top dead center) that is later than the timing at which compression self-ignition originally occurs, and combustion is established by ignition even in a cycle in which compression self-ignition ends. It becomes possible to prevent misfiring and deterioration of emission in the vicinity of the relatively high load operation region where the compression self-ignition operation is switched to the spark ignition operation.

  In the present embodiment, the number of spark plugs 34 is set to at least three including those laid out at the center of the cylinder 24. Therefore, in the present embodiment, when performing multipoint ignition, one spark plug 34 ignites the air-fuel mixture from the central portion of the cylinder 24, and the remaining spark plug 34 ignites the air-fuel mixture from other parts. In addition, the propagation of combustion becomes faster, and even when multi-point ignition is not executed, the use of the spark plug 34 disposed in the central portion makes it possible to burn the air-fuel mixture relatively quickly. .

  In the present embodiment, the engine 20 has the cylinder bore center Z of the cylinder 24 offset from the rotation center O of the crankshaft 21 to the right as viewed from the side in which the rotation direction of the crankshaft 21 is clockwise. For this reason, in this embodiment, the ascending / descending speed of the piston 26 is asymmetric with respect to the top dead center, and the descending speed of the piston 26 at the initial stage of the expansion stroke is relatively slow. For this reason, since the combustion speed is relatively faster than the descending speed of the piston 26, it is possible to maintain a good combustion environment, and the energy acting on the piston 26 is increased, so that the fuel consumption can be improved. become.

  Moreover, in this embodiment, the swirl production | generation on-off valve 43a is provided as a swirl production | generation means which produces | generates a swirl in a cylinder. For this reason, in this embodiment, turbulent energy can be effectively maintained up to the top dead center by the generation of swirl, and in addition to multipoint ignition, it can contribute to the improvement of the combustion speed during the expansion stroke. For this reason, even when performing multipoint ignition after top dead center, efficiency deterioration can be suppressed. Also, in the compression self-ignition operation region at the time of partial load, it is effective for mixing fuel and air, and in the spark ignition operation region, it is effective for suppressing knocking due to rapid combustion.

  Further, in the present embodiment, the closing timing of the intake valve 30 is a predetermined timing before the bottom dead center, and the first compression timing when the effective compression ratio is substantially equal to the expansion ratio and the effective compression ratio are smaller than the expansion ratio. The variable valve timing mechanism 40 that can be changed to the second timing is provided, and the control unit 100 controls the variable valve timing mechanism 40 so that the closing timing of the intake valve becomes the second timing at least when the engine is warm. To control. For this reason, in the present embodiment, in the operation region where high output of the engine 20 is required, the intake valve 30 is driven so that the effective compression ratio is substantially equal to the expansion ratio, thereby improving the output by the high compression ratio. On the other hand, at the time of warm start that requires a low starting torque, it is possible to drive the intake valve 30 so that the effective compression ratio is low, and to improve the startability by reducing the compression resistance.

  Thus, according to the present embodiment, there is a remarkable effect that knocking can be suppressed and fuel consumption can be improved while adopting as high a geometric compression ratio as possible.

  The above-described embodiments are merely preferred specific examples of the present invention, and the present invention is not limited to the above-described embodiments.

  For example, you may control the engine 20 shown in FIGS. 1-7 based on the characteristic shown in FIG.

  FIG. 11 is a characteristic diagram according to another embodiment of the present invention.

  As shown in the figure, in the partial load operation region D3, the external EGR system 60 is operated to introduce the external EGR into the cylinder, and in the relatively high load operation region of the EGR introduction region, multiple points are obtained after compression top dead center. The control unit 100 may control the external EGR system 60 and the spark plug 34 to fire.

  In this embodiment, since the geometric compression ratio of the engine 20 is high, knocking is likely to occur from the partial load operation region. However, the introduction of external EGR increases the specific heat and suppresses the temperature rise of the air-fuel mixture. Is suppressed. Furthermore, the rapid combustion effect of multi-point ignition is added, the knocking suppression effect is increased, the knocking suppression region can be expanded at the time of partial load, and the fuel efficiency can be improved at the same time.

  Moreover, in each embodiment mentioned above, as the fuel injection valve 32 and the ignition plug 34, you may employ | adopt the structure of FIG. 12, FIG.

  12 is a cross-sectional view of an engine 20 according to still another embodiment of the present invention, and FIG. 13 is an enlarged schematic view of a cylinder portion according to the embodiment of FIG.

  Referring to FIGS. 12 and 13, as the fuel injection valve 32, a center injection method arranged at the center portion of the cylinder 24 may be adopted. Furthermore, the number of spark plugs 34 is not limited to three, and for example, a configuration of four spark plugs 34 disposed between the intake ports 28 and 28 may be used.

  In the embodiment of FIGS. 12 and 13, fuel is injected from the central portion of the cylinder 24 and multipoint ignition is performed by the four spark plugs 34, so that the combustion speed can be further increased.

  Further, in each of the above-described embodiments, the means for generating the swirl is not limited to the above-described on-off valve 43a. For example, two intake valves 30 and two exhaust valves 31 are provided per cylinder, and one intake valve 30 is opened / closed. The operation may be stopped, and the exhaust valve positioned diagonally with respect to the intake valve that performs the valve opening operation may be configured to resume the valve operation.

  Further, as a method for obtaining the Miller cycle effect of blowing the burned gas back to the intake port 28, not only when the intake valve 30 is closed late, but also early opening that opens the intake valve 30 in the latter half of the exhaust stroke may be adopted.

  Further, in realizing multipoint ignition, in multipoint ignition at high load, the ignition timings of the spark plugs arranged in the central portion and the spark plugs arranged in the peripheral portion may be shifted.

  It goes without saying that various modifications can be made within the scope of the claims of the present invention.

It is a block diagram which shows schematic structure of the control apparatus which concerns on one Embodiment of this invention. 2 is a schematic cross-sectional view showing the structure of one cylinder of the four-cycle gasoline engine according to FIG. 1. It is a model figure of the piston in this embodiment. It is a graph showing the speed of the piston with respect to a crank angle. It is a model figure which shows the relationship between a crank angle and piston movement amount. 2 is a schematic plan view showing a cylinder 24 in an enlarged manner. It is explanatory drawing which shows the airflow of the combustion chamber which concerns on this embodiment, (A) has shown the compression stroke initial stage, (B) has shown the expansion stroke initial stage, respectively. It is a characteristic view which shows an example of the driving | operation area | region setting for performing control according to the driving | running state which concerns on embodiment of FIG. 6 is a timing chart showing intake valve opening timing and ignition timing for each characteristic. It is a characteristic view which shows the relationship between an engine load and an engine speed. It is a characteristic view concerning another embodiment of the present invention. It is sectional drawing of the engine which concerns on another embodiment of this invention. FIG. 13 is an enlarged schematic view of a cylinder portion according to the embodiment of FIG. 12.

Explanation of symbols

DESCRIPTION OF SYMBOLS 10 Control apparatus 20 4 cycle gasoline engine 21 Crankshaft 24 Cylinder 26a, 26b Squish area structure surface 26 Piston 27 Combustion chamber 28 Intake port 29 Exhaust port 30 Intake valve 31 Exhaust valve 32 Fuel injection valve 34 Spark plug 35 Ignition circuit 40 Variable valve Timing mechanism 43a Swirl generating on-off valve 46 Intake passage 50 Heater 60 External EGR system 100 Control unit O Rotation center S Offset amount SW1 Air flow sensor (an example of operation state detection means)
SW2 intake air temperature sensor (an example of operation state detection means)
SW3 Crank angle sensor (an example of operation state detection means)
SW4 engine water temperature sensor (an example of operation state detection means)
SW5 Oxygen concentration sensor (an example of operation state detection means)
SW6 Accelerator opening sensor (an example of operation state detection means)
Z Cylinder bore center

Claims (8)

A crankshaft that outputs torque, a piston that is coupled to the crankshaft, and a cylinder that partitions the combustion chamber of the air-fuel mixture in cooperation with the piston by fitting the piston in a reciprocating manner. In a control device for a four-cycle spark ignition engine for a vehicle,
Setting the engine's geometric compression ratio to 14 or more,
An operating state detecting means for detecting an operating state of the engine;
Multiple spark plugs per cylinder,
Control means for controlling the operation of the spark plug based on the detection of the operating state detection means, and in a high load operation area in the engine low speed area that is higher than the idle speed by a predetermined speed, multiple points after the compression top dead center A control device for a four-cycle spark ignition engine for a vehicle, wherein the control means controls the spark plug so as to fire. The control device for a four-cycle spark ignition engine for a vehicle according to claim 1,
An intake air heating means for heating the intake air is provided,
In the partial load operation region, the control means controls the intake air heating means so as to heat the in-cylinder intake air and perform the compression self-ignition operation. apparatus. The control device for a four-cycle spark ignition engine for a vehicle according to claim 2,
4. A control apparatus for a four-cycle spark ignition engine for a vehicle, wherein on the high load side of the partial load operation region, the control means controls the spark plug so that multipoint ignition is performed after compression self-ignition. The control device for a four-cycle spark ignition engine for a vehicle according to claim 1,
An external EGR device that recirculates a portion of the burned gas to the intake passage;
In the partial load operation region, the external EGR is introduced, and in the relatively high load operation region of the EGR introduction region, the control means controls the external EGR device and the spark plug so that multipoint ignition is performed after compression top dead center. A control device for a four-cycle spark ignition engine for vehicles. The control device for a four-cycle spark ignition engine for a vehicle according to any one of claims 1 to 4,
The control device for a four-cycle spark ignition engine for vehicles, wherein the number of the spark plugs is set to at least three including those laid out in the center of the cylinder. The control device for a four-cycle spark ignition engine for a vehicle according to any one of claims 1 to 5,
In the four-cycle spark ignition engine for a vehicle, the center of the cylinder bore of the cylinder is offset from the center of rotation of the crankshaft to the right side when viewed from the side in which the rotation direction of the crankshaft is clockwise. Control device. The control device for a four-cycle spark ignition engine for a vehicle according to any one of claims 1 to 6,
A control apparatus for a four-cycle spark ignition engine for a vehicle, comprising swirl generating means for generating swirl in a cylinder. The control device for a four-cycle spark ignition engine for a vehicle according to any one of claims 1 to 7,
The closing timing of the intake valve is changed to a first timing at which the effective compression ratio is substantially equal to the expansion ratio and a second timing at which the effective compression ratio is smaller than the expansion ratio. Possible variable valve timing mechanism,
The four-cycle spark ignition type for a vehicle, wherein the control means controls the variable valve timing mechanism so that the closing timing of the intake valve becomes the second timing at least when the engine is warm. Engine control device.

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