JP2011117418A - Spark ignition internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To inhibit generation of noise from an intake passage. <P>SOLUTION: A spark ignition internal combustion engine includes a variable compression ratio mechanism A variable in a mechanical compression ratio, a variable valve timing mechanism B capable of controlling valve close timing of an intake valve 7, and a throttle valve 17 disposed in an engine intake passage. Noise caused by intake air interference increases when retardation quantity of valve close timing of the intake valve 7 and opening of the throttle valve 17 become large. Consequently, opening of the throttle valve 7 is reduced and valve close timing of the intake valve 7 is advanced when the engine is in an operation state where the noise is to be reduced. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

  The present invention relates to a spark ignition internal combustion engine.

  Equipped with a variable compression ratio mechanism that can change the mechanical compression ratio, a variable valve timing mechanism that can control the closing timing of the intake valve, and a throttle valve that is arranged in the engine intake passage. On the side, the intake air amount is controlled by controlling the retard amount of the closing timing of the intake valve while the throttle valve is fully opened, and the closing timing of the intake valve is most retarded during engine low load operation And an internal combustion engine in which the amount of intake air is controlled by a throttle valve is known (see, for example, Patent Document 1).

JP 2007-303423 A

  By the way, in an internal combustion engine, when the piston descends during the intake stroke, negative pressure is generated in the intake port, and this negative pressure is reflected at the opening end to the surge tank and returns to the intake port again in the form of positive pressure. . On the other hand, when the closing timing of the intake valve is retarded as in the above-described internal combustion engine, the air compressed by the piston is pushed into the intake port, so that positive pressure is generated in the intake port. This positive pressure is superimposed on the positive pressure returned from the open end to the surge tank, and thus a large positive pressure is generated in the intake port. As a result, intense pressure pulsation occurs in the intake passage due to such intake air interference.

  However, in the above-described internal combustion engine, when such pressure pulsation occurs, the throttle valve is kept fully open. As a result, this pressure pulsation propagates to the air intake without being obstructed by the throttle valve, thus causing a problem that noise is generated. Further, when an intake air amount detector is provided in the intake passage, there is a problem that a measurement error is caused by this pressure pulsation.

  In order to solve the above problem, according to the present invention, a variable compression ratio mechanism capable of changing the mechanical compression ratio, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and the engine intake passage are disposed. In a spark ignition internal combustion engine equipped with a throttle valve, when the retard amount of the closing timing of the intake valve and the opening of the throttle valve are increased, noise due to intake interference increases, and this noise must be reduced. Sometimes, the opening of the throttle valve is reduced and the closing timing of the intake valve is advanced.

  Since the opening degree of the throttle valve is reduced when the engine operation state is an operation state where noise should be reduced, propagation of pressure pulsation is suppressed. At this time, the closing timing of the intake valve is advanced, so that the positive pressure generated in the intake port is weakened, and thus the pressure pulsation is weakened. As a result, generation of noise can be suppressed.

1 is an overall view of a spark ignition internal combustion engine. It is a disassembled perspective view of a variable compression ratio mechanism. 1 is a schematic side sectional view of an internal combustion engine. It is a figure which shows a variable valve timing mechanism. It is a figure which shows the lift amount of an intake valve and an exhaust valve. It is a figure for demonstrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. It is a figure which shows the relationship between theoretical thermal efficiency and an expansion ratio. It is a figure for demonstrating a normal cycle and a super-high expansion ratio cycle. It is a figure which shows changes, such as a mechanical compression ratio according to an engine load. It is a figure for demonstrating the pressure in an intake port. It is a figure which shows the pressure pulsation in an intake duct upstream end. It is a figure which shows changes, such as a mechanical compression ratio according to an engine load. It is a flowchart for performing operation control. It is a figure which shows maps, such as valve closing timing of an intake valve. It is a figure which shows maps, such as valve closing timing of an intake valve. It is a general view of another Example of a spark ignition type internal combustion engine. It is a figure for demonstrating control with respect to intake pipe length.

FIG. 1 shows a side sectional view of a spark ignition type internal combustion engine.
Referring to FIG. 1, 1 is a crankcase, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is a spark plug disposed at the center of the top surface of the combustion chamber 5, and 7 is intake air. 8 is an intake port, 9 is an exhaust valve, and 10 is an exhaust port. The intake port 8 is connected to a surge tank 12 via an intake branch pipe 11, and a fuel injection valve 13 for injecting fuel into the corresponding intake port 8 is arranged in each intake branch pipe 11. The fuel injection valve 13 may be arranged in each combustion chamber 5 instead of being attached to each intake branch pipe 11.

  The surge tank 12 is connected to an air cleaner 15 via an intake duct 14, and a throttle valve 17 driven by an actuator 16 and an intake air amount detector 18 using, for example, heat rays are arranged in the intake duct 14. On the other hand, the exhaust port 10 is connected to a catalytic converter 20 containing, for example, a three-way catalyst via an exhaust manifold 19, and an air-fuel ratio sensor 21 is disposed in the exhaust manifold 19.

  On the other hand, in the embodiment shown in FIG. 1, the piston 4 is positioned at the compression top dead center by changing the relative position of the crankcase 1 and the cylinder block 2 in the cylinder axial direction at the connecting portion between the crankcase 1 and the cylinder block 2. There is provided a variable compression ratio mechanism A capable of changing the volume of the combustion chamber 5 at the time, and further an actual compression action start timing changing mechanism B capable of changing the actual start time of the compression action. In the embodiment shown in FIG. 1, the actual compression action start timing changing mechanism B is composed of a variable valve timing mechanism capable of controlling the closing timing of the intake valve 7.

  The electronic control unit 30 is composed of a digital computer, and is connected to each other by a bidirectional bus 31. A ROM (Read Only Memory) 32, a RAM (Random Access Memory) 33, a CPU (Microprocessor) 34, an input port 35 and an output port 36. It comprises. The output signal of the intake air amount detector 18 and the output signal of the air-fuel ratio sensor 21 are input to the input port 35 via corresponding AD converters 37, respectively. A load sensor 41 that generates an output voltage proportional to the depression amount L of the accelerator pedal 40 is connected to the accelerator pedal 40, and the output voltage of the load sensor 41 is input to the input port 35 via the corresponding AD converter 37. Is done. Further, a crank angle sensor 42 that generates an output pulse every time the crankshaft rotates, for example, 30 ° is connected to the input port 35. On the other hand, the output port 36 is connected to the spark plug 6, the fuel injection valve 13, the throttle valve driving actuator 16, the variable compression ratio mechanism A, and the variable valve timing mechanism B through corresponding drive circuits 38.

  2 is an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 1, and FIG. 3 is a side sectional view of the internal combustion engine schematically shown. Referring to FIG. 2, a plurality of protrusions 50 spaced from each other are formed below both side walls of the cylinder block 2, and cam insertion holes 51 each having a circular cross section are formed in each protrusion 50. Has been. On the other hand, a plurality of protrusions 52 are formed on the upper wall surface of the crankcase 1 so as to be fitted between the corresponding protrusions 50 spaced apart from each other. Cam insertion holes 53 each having a circular cross section are formed.

  As shown in FIG. 2, a pair of camshafts 54 and 55 are provided, and on each camshaft 54 and 55, a circular cam 56 is rotatably inserted into each cam insertion hole 51. It is fixed. These circular cams 56 are coaxial with the rotational axes of the camshafts 54 and 55. On the other hand, an eccentric shaft 57 arranged eccentrically with respect to the rotation axis of each camshaft 54, 55 extends between the circular cams 56 as shown by hatching in FIG. A cam 58 is eccentrically mounted for rotation. As shown in FIG. 2, the circular cams 58 are disposed between the circular cams 56, and the circular cams 58 are rotatably inserted into the corresponding cam insertion holes 53.

  When the circular cams 56 fixed on the camshafts 54 and 55 are rotated in opposite directions as indicated by solid arrows in FIG. 3A from the state shown in FIG. In order to move toward the lower center, the circular cam 58 rotates in the opposite direction to the circular cam 56 in the cam insertion hole 53 as shown by the broken arrow in FIG. 3A, and is shown in FIG. As described above, when the eccentric shaft 57 moves to the lower center, the center of the circular cam 58 moves below the eccentric shaft 57.

  3A and 3B, the relative positions of the crankcase 1 and the cylinder block 2 are determined by the distance between the center of the circular cam 56 and the center of the circular cam 58. The cylinder block 2 moves away from the crankcase 1 as the distance between the center and the center of the circular cam 58 increases. When the cylinder block 2 moves away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is positioned at the compression top dead center. Therefore, by rotating the camshafts 54 and 55, the piston 4 is compressed at the top dead center. The volume of the combustion chamber 5 when it is located at can be changed.

  As shown in FIG. 2, in order to rotate the camshafts 54 and 55 in the opposite directions, a pair of worm gears 61 and 62 having opposite spiral directions are attached to the rotation shaft of the drive motor 59, respectively. Gears 63 and 64 meshing with the worm gears 61 and 62 are fixed to the end portions of the camshafts 54 and 55, respectively. In this embodiment, by driving the drive motor 59, the volume of the combustion chamber 5 when the piston 4 is located at the compression top dead center can be changed over a wide range. The variable compression ratio mechanism A shown in FIGS. 1 to 3 shows an example, and any type of variable compression ratio mechanism can be used.

  On the other hand, FIG. 4 shows the variable valve timing mechanism B attached to the end of the camshaft 70 for driving the intake valve 7 in FIG. Referring to FIG. 4, the variable valve timing mechanism B includes a timing pulley 71 that is rotated in the direction of an arrow by a crankshaft of an engine via a timing belt, a cylindrical housing 72 that rotates together with the timing pulley 71, an intake valve A rotating shaft 73 that rotates together with the driving camshaft 70 and is rotatable relative to the cylindrical housing 72, and a plurality of partition walls 74 that extend from the inner peripheral surface of the cylindrical housing 72 to the outer peripheral surface of the rotating shaft 73. And a vane 75 extending from the outer peripheral surface of the rotating shaft 73 to the inner peripheral surface of the cylindrical housing 72 between the partition walls 74, and an advance hydraulic chamber 76 on each side of each vane 75. A retarding hydraulic chamber 77 is formed.

  The hydraulic oil supply control to the hydraulic chambers 76 and 77 is performed by a hydraulic oil supply control valve 78. The hydraulic oil supply control valve 78 includes hydraulic ports 79 and 80 connected to the hydraulic chambers 76 and 77, a hydraulic oil supply port 82 discharged from the hydraulic pump 81, a pair of drain ports 83 and 84, And a spool valve 85 for controlling communication between the ports 79, 80, 82, 83, and 84.

  When the cam phase of the intake valve driving camshaft 70 should be advanced, the spool valve 85 is moved to the right in FIG. 4 and the hydraulic oil supplied from the supply port 82 is advanced via the hydraulic port 79. The hydraulic oil in the retard hydraulic chamber 77 is discharged from the drain port 84 while being supplied to the hydraulic chamber 76. At this time, the rotary shaft 73 is rotated relative to the cylindrical housing 72 in the direction of the arrow.

  On the other hand, when the cam phase of the intake valve driving camshaft 70 should be retarded, the spool valve 85 is moved to the left in FIG. 4, and the hydraulic oil supplied from the supply port 82 causes the hydraulic port 80 to move. The hydraulic oil in the advance hydraulic chamber 76 is discharged from the drain port 83 while being supplied to the retard hydraulic chamber 77. At this time, the rotating shaft 73 is rotated relative to the cylindrical housing 72 in the direction opposite to the arrow.

  If the spool valve 85 is returned to the neutral position shown in FIG. 4 while the rotation shaft 73 is rotated relative to the cylindrical housing 72, the relative rotation operation of the rotation shaft 73 is stopped, and the rotation shaft 73 is The relative rotation position at that time is held. Therefore, the variable valve timing mechanism B can advance and retard the cam phase of the intake valve driving camshaft 70 by a desired amount.

  In FIG. 5, the solid line shows the time when the cam phase of the intake valve driving camshaft 70 is advanced the most by the variable valve timing mechanism B, and the broken line shows the cam phase of the intake valve driving camshaft 70 being the most advanced. It shows when it is retarded. Therefore, the valve opening period of the intake valve 7 can be arbitrarily set between the range indicated by the solid line and the range indicated by the broken line in FIG. 5, and therefore the closing timing of the intake valve 7 is also the range indicated by the arrow C in FIG. Any crank angle can be set.

  The variable valve timing mechanism B shown in FIG. 1 and FIG. 4 shows an example. For example, the variable valve timing that can change only the closing timing of the intake valve while keeping the opening timing of the intake valve constant. Various types of variable valve timing mechanisms, such as mechanisms, can be used.

  Next, the meanings of terms used in the present application will be described with reference to FIG. 6 (A), (B), and (C) show an engine having a combustion chamber volume of 50 ml and a piston stroke volume of 500 ml for the sake of explanation. , (B), (C), the combustion chamber volume represents the volume of the combustion chamber when the piston is located at the compression top dead center.

  FIG. 6A explains the mechanical compression ratio. The mechanical compression ratio is a value mechanically determined only from the stroke volume of the piston and the combustion chamber volume during the compression stroke, and this mechanical compression ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 6A, this mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11.

  FIG. 6B describes the actual compression ratio. This actual compression ratio is a value determined from the actual piston stroke volume and the combustion chamber volume from when the compression action is actually started until the piston reaches top dead center, and this actual compression ratio is (combustion chamber volume + actual (Stroke volume) / combustion chamber volume. That is, as shown in FIG. 6B, even if the piston starts to rise in the compression stroke, the compression operation is not performed while the intake valve is open, and the actual compression is performed from the time when the intake valve is closed. The action begins. Therefore, the actual compression ratio is expressed as described above using the actual stroke volume. In the example shown in FIG. 6B, the actual compression ratio is (50 ml + 450 ml) / 50 ml = 10.

  FIG. 6C illustrates the expansion ratio. The expansion ratio is a value determined from the stroke volume of the piston and the combustion chamber volume during the expansion stroke, and this expansion ratio is expressed by (combustion chamber volume + stroke volume) / combustion chamber volume. In the example shown in FIG. 6C, this expansion ratio is (50 ml + 500 ml) / 50 ml = 11.

  Next, an ultra-high expansion ratio cycle used in the present invention will be described with reference to FIGS. FIG. 7 shows the relationship between the theoretical thermal efficiency and the expansion ratio, and FIG. 8 shows a comparison between a normal cycle and an ultrahigh expansion ratio cycle that are selectively used according to the load in the present invention.

  FIG. 8A shows a normal cycle when the intake valve closes near the bottom dead center and the compression action by the piston is started from the vicinity of the intake bottom dead center. In the example shown in FIG. 8A, the combustion chamber volume is 50 ml, and the stroke volume of the piston is 500 ml, as in the examples shown in FIGS. 6A, 6B, and 6C. As can be seen from FIG. 8A, in a normal cycle, the mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11, the actual compression ratio is almost 11, and the expansion ratio is also (50 ml + 500 ml) / 50 ml = 11. That is, in a normal internal combustion engine, the mechanical compression ratio, the actual compression ratio, and the expansion ratio are substantially equal.

  The solid line in FIG. 7 shows the change in theoretical thermal efficiency when the actual compression ratio and the expansion ratio are substantially equal, that is, in a normal cycle. In this case, it can be seen that the theoretical thermal efficiency increases as the expansion ratio increases, that is, as the actual compression ratio increases. Therefore, in order to increase the theoretical thermal efficiency in a normal cycle, it is only necessary to increase the actual compression ratio. However, the actual compression ratio can only be increased to a maximum of about 12 due to the restriction of the occurrence of knocking at the time of engine high load operation, and thus the theoretical thermal efficiency cannot be sufficiently increased in a normal cycle.

  On the other hand, in this situation, the present inventor has studied to increase the theoretical thermal efficiency by strictly dividing the mechanical compression ratio and the actual compression ratio, and as a result, the theoretical thermal efficiency is governed by the expansion ratio, and the theoretical thermal efficiency The actual compression ratio has been found to have little effect. That is, if the actual compression ratio is increased, the explosive force is increased, but a large amount of energy is required for compression. Thus, even if the actual compression ratio is increased, the theoretical thermal efficiency is hardly increased.

  On the other hand, when the expansion ratio is increased, the period during which the pressing force acts on the piston during the expansion stroke becomes longer, and thus the period during which the piston applies the rotational force to the crankshaft becomes longer. Therefore, the larger the expansion ratio, the higher the theoretical thermal efficiency. The broken line ε = 10 in FIG. 7 indicates the theoretical thermal efficiency when the expansion ratio is increased with the actual compression ratio fixed at 10. Thus, the theoretical thermal efficiency when the expansion ratio is increased while the actual compression ratio is maintained at a low value, and the theoretical thermal efficiency when the actual compression ratio is increased with the expansion ratio as shown by the solid line in FIG. It can be seen that there is no significant difference from the amount of increase.

  Thus, if the actual compression ratio is maintained at a low value, knocking does not occur. Therefore, if the expansion ratio is increased while the actual compression ratio is maintained at a low value, the theoretical thermal efficiency is reduced while preventing the occurrence of knocking. Can be greatly increased. FIG. 8B shows an example where the variable compression ratio mechanism A and variable valve timing mechanism B are used to increase the expansion ratio while maintaining the actual compression ratio at a low value.

  Referring to FIG. 8B, in this example, the variable compression ratio mechanism A reduces the combustion chamber volume from 50 ml to 20 ml. On the other hand, the variable valve timing mechanism B delays the closing timing of the intake valve until the actual piston stroke volume is reduced from 500 ml to 200 ml. As a result, in this example, the actual compression ratio is (20 ml + 200 ml) / 20 ml = 11, and the expansion ratio is (20 ml + 500 ml) / 20 ml = 26. In the normal cycle shown in FIG. 8A, the actual compression ratio is almost 11 and the expansion ratio is 11, as described above. Compared with this case, only the expansion ratio is shown in FIG. 8B. It can be seen that it has been increased to 26. This is why it is called an ultra-high expansion ratio cycle.

  Generally speaking, in an internal combustion engine, the lower the engine load, the worse the thermal efficiency. Therefore, in order to improve the thermal efficiency during engine operation, that is, to improve fuel efficiency, it is necessary to improve the thermal efficiency when the engine load is low. Necessary. On the other hand, in the ultra-high expansion ratio cycle shown in FIG. 8B, since the actual piston stroke volume during the compression stroke is reduced, the amount of intake air that can be sucked into the combustion chamber 5 is reduced. The expansion ratio cycle can only be adopted when the engine load is relatively low. Therefore, in the present invention, when the engine load is relatively low, the super high expansion ratio cycle shown in FIG. 8B is used, and during the high engine load operation, the normal cycle shown in FIG. 8A is used.

Next, the basic operation control in general will be described with reference to FIG.
FIG. 9 shows changes in the mechanical compression ratio, the expansion ratio, the closing timing of the intake valve 7, the actual compression ratio, the intake air amount, and the opening degree of the throttle valve 17 according to the engine load at a certain engine speed. ing. 9 shows that the average air-fuel ratio in the combustion chamber 5 is the output signal of the air-fuel ratio sensor 21 so that unburned HC, CO and NO _x in the exhaust gas can be simultaneously reduced by the three-way catalyst in the catalytic converter 20. This shows a case where feedback control is performed to the stoichiometric air-fuel ratio based on the above.

  As described above, the normal cycle shown in FIG. 8 (A) is executed during engine high load operation. Accordingly, as shown in FIG. 9, the expansion ratio is low because the mechanical compression ratio is lowered at this time, and the closing timing of the intake valve 7 is advanced as shown by the solid line in FIG. 5 as shown in FIG. Yes. At this time, the intake air amount is large, and at this time, the opening degree of the throttle valve 17 is kept fully open.

  On the other hand, as shown in FIG. 9, when the engine load becomes low, the closing timing of the intake valve 7 is retarded so as to reduce the intake air amount. Further, at this time, as shown in FIG. 9, the mechanical compression ratio is increased as the engine load is lowered so that the actual compression ratio is kept substantially constant. Therefore, the expansion ratio is also increased as the engine load is lowered. At this time, the throttle valve 17 is kept fully open, and therefore the amount of intake air supplied into the combustion chamber 5 is controlled by changing the closing timing of the intake valve 7 regardless of the throttle valve 17. Has been.

  As described above, when the engine load is reduced from the engine high load operation state, the mechanical compression ratio is increased as the intake air amount is decreased while the actual compression ratio is substantially constant. That is, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is decreased in proportion to the decrease in the intake air amount. Therefore, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the intake air amount. At this time, in the example shown in FIG. 9, the air-fuel ratio in the combustion chamber 5 is the stoichiometric air-fuel ratio, so the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is proportional to the fuel amount. Will change.

  When the engine load is further reduced, the mechanical compression ratio is further increased, and when the engine load is lowered to a medium load L slightly close to a low load, the mechanical compression ratio reaches a limit mechanical compression ratio that is a structural limit of the combustion chamber 5. When the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is maintained at the limit mechanical compression ratio in a region where the load is lower than the engine load L when the mechanical compression ratio reaches the limit mechanical compression ratio. Accordingly, the mechanical compression ratio is maximized and the expansion ratio is maximized at the time of low engine load operation and low engine load operation, that is, at the engine low load operation side. In other words, the mechanical compression ratio is maximized so that the maximum expansion ratio is obtained on the engine low load operation side.

  On the other hand, in the embodiment shown in FIG. 9, when the engine load decreases to L, the closing timing of the intake valve 7 becomes the limit closing timing that can control the amount of intake air supplied into the combustion chamber 5. When the closing timing of the intake valve 7 reaches the limit closing timing, the closing timing of the intake valve 7 is in a region where the load is lower than the engine load L when the closing timing of the intake valve 7 reaches the closing timing. It is held at the limit closing timing.

  When the closing timing of the intake valve 7 is held at the limit closing timing, the amount of intake air can no longer be controlled by changing the closing timing of the intake valve 7. In the embodiment shown in FIG. 9, at this time, that is, in a region where the load is lower than the engine load L when the valve closing timing of the intake valve 7 reaches the limit valve closing timing, the intake valve 7 is supplied into the combustion chamber 5 by the throttle valve 17. The amount of intake air to be controlled is controlled. At this time, the opening degree of the throttle valve 17 is made smaller as the engine load decreases.

  As described above, the expansion ratio is 26 in the ultra-high expansion ratio cycle shown in FIG. The higher the expansion ratio, the better. However, as can be seen from FIG. 7, a considerably high theoretical thermal efficiency can be obtained if it is 20 or more with respect to the practically usable lower limit actual compression ratio ε = 5. Therefore, in the present invention, the variable compression ratio mechanism A is formed so that the expansion ratio is 20 or more.

  Next, the pressure change in the intake port 8 will be described with reference to FIG. The vertical axis in FIG. 10 indicates the pressure in the intake port 8, and the horizontal axis in FIG. 10 indicates the crank angle from when the intake stroke is started. Further, the solid line in FIG. 10 schematically shows the pressure generated in the intake port 8 during the opening period and closing of the intake valve 7. That is, as shown by the solid line in FIG. 10, when the intake stroke starts and the piston 4 descends, negative pressure is generated in the intake port 8. Next, when the piston 4 starts to rise past the intake bottom dead center (BDC), the air compressed by the piston 4 is pushed into the intake port 8, and a positive pressure is generated in the intake port 8.

On the other hand, in FIG. 10, IC ₁ indicates the closing timing of the intake valve 7 at the time of high load operation, that is, the closing timing of the intake valve 7 at the most advanced angle, and IC ₂ is the limit closing timing, that is, The closing timing of the intake valve 7 when it is most retarded is shown. As indicated by IC ₁ , when the closing timing of the intake valve 7 is most advanced, the intake valve 7 is closed as soon as the compression action of the intake air by the piston 4 is started. A large positive pressure is not generated.

On the other hand, as indicated by IC ₂ , when the closing timing of the intake valve 7 is most retarded, the intake valve 7 is closed after the intake air is sufficiently compressed by the piston 4. A large positive pressure is generated in the port 8. In this case, roughly speaking, as the retard amount of the closing timing of the intake valve 7 increases, a larger positive pressure is generated in the intake port 8.

  As described above, when the piston 4 is lowered during the intake stroke, a negative pressure is generated in the intake port 8. This negative pressure is reflected at the opening end to the surge tank 12, and returns to the intake port 8 again in the form of a positive pressure. On the other hand, as described above, when the closing timing of the intake valve 7 is retarded, the air compressed by the piston 4 is pushed into the intake port 8, so that a positive pressure is generated in the intake port 8. When this positive pressure is superimposed on the positive pressure returned from the opening end to the surge tank 12, a large positive pressure is generated in the intake port 8, and as a result, intense pressure pulsation is generated in the intake passage due to intake air interference. It will be.

  By the way, when the throttle valve 17 is closed when such pressure pulsation due to intake interference occurs, the propagation of the pressure pulsation upstream is inhibited by the throttle valve 17, and thus the throttle valve 17 Propagation of pressure pulsations upstream is suppressed. On the other hand, if the throttle valve 17 is kept fully open when such pressure pulsation occurs, the pressure pulsation is not obstructed by the throttle valve 17, that is, the intake duct 14. Propagates to the upstream end. However, when the pressure pulsation is propagated to the air intake port as described above, noise is generated, and the intake air amount detector 18 causes a measurement error.

The intense pressure pulsation occurs when the positive pressure generated in the intake port 8 due to the pushing action of the compressed air by the piston 4 overlaps with the positive pressure returned from the opening end to the surge tank 12, Such heavy pressure action occurs in a specific engine speed range determined from the reciprocation time of the pressure wave between the intake port 8 and the open end to the surge tank 12. In the internal combustion engine shown in FIG. 1, this rotational speed range is between N ₁ and N ₂ in FIG. FIG. 11 shows the magnitude of pressure fluctuation generated at the air intake port due to pressure pulsation due to intake air interference when the throttle valve 17 is fully open.

  As can be understood from the above description, when the retard amount of the intake valve 7 is increased, the pressure pulsation becomes intense, and as the opening of the throttle valve 17 is increased, the pressure pulsation reaching the air intake becomes stronger. That is, when the retard amount of the closing timing of the intake valve 7 and the opening degree of the throttle valve 17 are increased, noise based on intake interference increases. Therefore, in the present invention, the opening degree of the throttle valve 17 is reduced and the closing timing of the intake valve 7 is advanced in an operating state in which noise should be reduced. Next, this will be described with reference to FIG.

  FIG. 12 is a view similar to FIG. 9 and shows changes in the mechanical compression ratio and the like according to the load. In FIG. 12, broken lines indicate changes in the mechanical compression ratio and the like during the basic operation control shown in FIG. As shown in FIG. 12, according to the present invention, when the engine is operating near the load L where noise is to be reduced, the opening of the throttle valve 17 is lowered and the closing timing of the intake valve 7 is advanced. Horned. Thus, when the opening degree of the throttle valve 17 is reduced, the propagation of pressure pulsation is suppressed. On the other hand, when the closing timing of the intake valve 7 is advanced, if it occurs in the intake port 8, the positive pressure is weakened, and thus the pressure pulsation is weakened. As a result, noise generation can be suppressed.

  Further, the advancement of the closing timing of the intake valve 7 when the opening of the throttle valve 17 is reduced is also to satisfy the required intake air amount. Further, as shown in FIG. 12, when the closing timing of the intake valve 7 is advanced, the mechanical compression is performed so that the compression end pressure and the compression end temperature do not increase excessively, that is, to maintain the actual compression ratio constant. The ratio is lowered.

As described above, noise is generated when the retard amount of the closing timing of the intake valve 7 is large and the opening degree of the throttle valve 17 is large, that is, in the vicinity of the load L in FIG. Therefore, in the embodiment according to the present invention, the opening degree of the throttle valve 17 is lowered in the load region between L ₁ and L ₂ including the load L in FIG. 12, the closing timing of the intake valve 7 is advanced, and the mechanical compression ratio is set. I try to lower it. Incidentally, the opening degree of the throttle valve 17 in the embodiment shown in FIG. 12, the closing timing and the mechanical compression ratio of the intake valve 7 is changed linearly allowed according to the change of the load between the load L ₁ and the load L ₂ It is done.

That is, in the embodiment according to the present invention, the load region between L ₁ and L ₂ and the rotation speed region between N ₁ and N ₂ where the noise increases as the operating state in which noise should be reduced are obtained in advance, and the engine load Is within the load range between L ₁ and L ₂ and the engine speed is within the speed range between N ₁ and N ₂ , the opening of the throttle valve 17 is reduced and the intake valve 7 is closed. The time is advanced.

In this case, the load region between L ₁ and L ₂ is a load region that spreads to the high load side and the low load side around the load L at which the closing timing of the intake valve 7 becomes the limit closing timing.

Next, the operation control routine shown in FIG. 13 will be described.
Referring to FIG. 13, first, at step 100, it is judged if the engine load is within the load region between L ₁ and L ₂ . When the engine load is not within the load region between L ₁ and L ₂ , the routine proceeds to step 102. On the other hand, when the engine load is within the load range between L ₁ and L ₂ , the routine proceeds to step 101 where it is judged if the engine speed is within the range between N ₁ and N ₂ . When the engine speed is not within the speed range between N ₁ and N ₂ , the routine proceeds to step 102.

  From step 102 to step 104, the basic operation control shown in FIG. 9 is performed. That is, at step 102, the closing timing IC of the intake valve 7 is calculated from the map shown in FIG. That is, when the basic operation control as shown in FIG. 9 is performed, the closing timing IC of the intake valve 7 necessary for supplying the required intake air amount into the combustion chamber 5 is determined by the engine load L and the engine speed. As a function of the number N, it is stored in advance in the ROM 32 in the form of a map as shown in FIG. 14A, and the valve closing timing IC of the intake valve 7 is calculated from this map.

  Next, at step 103, a mechanical compression ratio CR necessary for making the actual compression ratio constant is calculated. Next, at step 104, the opening degree θ of the throttle valve 17 when the basic operation control as shown in FIG. 9 is performed is calculated. The opening θ of the throttle valve 17 is stored in advance in the ROM 32 as a function of the engine load L and the engine speed N in the form of a map as shown in FIG. Next, at step 105, the variable compression ratio mechanism A is controlled so that the mechanical compression ratio becomes the mechanical compression ratio CR, and the variable valve timing mechanism B is controlled so that the closing timing of the intake valve 7 becomes the closing timing IC, The throttle valve 17 is controlled so that the opening degree of the throttle valve 17 becomes the opening degree θ.

On the other hand, when it is determined in step 101 that the engine speed is within the engine speed range between N ₁ and N ₂ , the routine proceeds to step 106 where operation control for noise control shown in FIG. 12 is performed. That is, at step 106, the closing timing IC ′ of the intake valve 7 is calculated from the map shown in FIG. That is, when the operation control for noise suppression shown in FIG. 12 is performed, the closing timing IC ′ of the intake valve 7 necessary for supplying the required intake air amount into the combustion chamber 5 is the engine load L and the engine. As a function of the rotational speed N, it is stored in advance in the ROM 32 in the form of a map as shown in FIG. 15A, and the closing timing IC ′ of the intake valve 7 is calculated from this map.

  Next, at step 107, a mechanical compression ratio CR 'necessary for making the actual compression ratio constant is calculated. Next, at step 108, the opening degree θ 'of the throttle valve 17 when the operation control for noise suppression shown in FIG. 12 is performed is calculated. The opening θ ′ of the throttle valve 17 is stored in advance in the ROM 32 as a function of the engine load L and the engine speed N in the form of a map as shown in FIG. Next, at step 105, the variable compression ratio mechanism A is controlled so that the mechanical compression ratio becomes the mechanical compression ratio CR ′, and the variable valve timing mechanism B is controlled so that the closing timing of the intake valve 7 becomes the closing timing IC ′. Then, the throttle valve 17 is controlled so that the opening degree of the throttle valve 17 becomes the opening degree θ ′.

  FIG. 16 shows another embodiment. In this embodiment, the intake branch pipe 11 of each cylinder extends so as to surround the surge tank 12, and intake air flows into the surge tank 12 from the intake inlet 22. In the middle of each intake branch pipe 11, it communicates with the inside of the surge tank 22 via the communication port 23, and a switching valve 24 for changing the intake pipe length is arranged in each intake branch pipe 11. When the switching valve 24 opens the communication port 23 and shuts off the corresponding intake branch pipe 11 as shown by a solid line in FIG. 16, the intake pipe length becomes short. As shown by the broken line in FIG. When the communication port 23 is closed, the length of the intake pipe becomes longer.

In this embodiment, the engine is normally operated by the switching action of the switching valve 24 so that the pressure wave reflected at the opening end to the surge tank 12 returns into the intake port 8 in the form of positive pressure when the intake valve 7 is opened. When the engine speed is high, the intake branch pipe is shortened, and when the engine speed is low, the intake branch pipe is lengthened. However, on the engine low load side where the valve closing period of the intake valve 7 is most retarded, there is a risk that severe pressure pulsation may occur due to the above-described heavy pressure action in the intake port 8 on the engine low rotation side. Therefore, in this embodiment, as shown in FIG. 17, in order to shift the arrival timing of the reflected wave, the engine load is on the engine low load side lower than L ₀ and the engine speed is lower than N _0. In some cases, the intake pipe length is shortened.

DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 3 Cylinder head 4 Piston 5 Combustion chamber 7 Intake valve 17 Throttle valve 70 Intake valve drive camshaft A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (4)

  In a spark ignition internal combustion engine comprising a variable compression ratio mechanism capable of changing the mechanical compression ratio, a variable valve timing mechanism capable of controlling the closing timing of the intake valve, and a throttle valve disposed in the engine intake passage, When the retard amount of the closing timing of the intake valve and the opening of the throttle valve are increased, noise due to intake interference increases, and when the operating state is to reduce the noise, the opening of the throttle valve is reduced and the intake valve A spark ignition internal combustion engine that advances the valve closing timing.   The spark ignition type internal combustion engine according to claim 1, wherein the mechanical compression ratio is lowered in an operating state in which the noise is to be reduced.   A load region and a rotational speed region where the noise increases as an operating state in which the noise is to be reduced are obtained in advance, and the throttle is set when the engine load is within the load region and the engine rotational speed is within the rotational speed region. The spark ignition internal combustion engine according to claim 1 or 2, wherein the opening degree of the valve is reduced and the closing timing of the intake valve is advanced.   A limit valve closing timing capable of controlling the intake air amount with respect to the valve closing timing of the intake valve is set in advance, and the load region is centered on a load at which the valve closing timing of the intake valve becomes the limit valve closing timing. The spark ignition type internal combustion engine according to claim 3, wherein the spark ignition type internal combustion engine is a load region extending to a high load side and a low load side.

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