JP5585526B2 - Spark ignition internal combustion engine - Google Patents

Description

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

  In order to reduce the pumping loss, it has been proposed to control the intake air amount by changing the closing timing of the intake valve by a variable valve timing mechanism without providing the throttle valve or by fully opening the throttle valve. (See Patent Document 1).

  In a spark ignition internal combustion engine having a pair of intake ports, the intake air amount is controlled by changing the closing timing of the two intake valves in the same manner.

JP2007-303423 JP2010-121617A JP 2006-161633 A JP 11-093712 A JP 07-180559 A

  In order to realize good combustion with a high combustion speed by generating turbulence in the cylinder, it is preferable to generate a strong air flow such as a swirl flow or a tumble flow in the cylinder during the intake stroke. For this purpose, in an internal combustion engine having a pair of intake ports, it is preferable that one of the intake ports be an airflow generation port for generating an airflow in the cylinder. Since the airflow generation port inevitably has a small flow coefficient, the other intake port is a straight port having a large flow coefficient in order to prevent intake shortage.

  In such a spark ignition internal combustion engine having a pair of intake ports, when the intake air amount is controlled by changing the closing timings of the two intake valves in the same manner, there remains room for improving combustion. .

  Accordingly, an object of the present invention is to improve combustion in a spark ignition internal combustion engine that includes an air flow generation port and a straight port and controls the intake air amount by changing the closing timing of the intake valve as compared with the conventional one. That is.

A spark ignition internal combustion engine according to a first aspect of the present invention includes a variable compression ratio mechanism and an intake port including an airflow generation port and a straight port , and changes an intake valve closing timing to reduce an intake air amount. In the spark ignition internal combustion engine to be controlled, when reducing the intake air amount, the opening of the second intake valve for the straight port is not changed without changing the opening timing and the closing timing of the first intake valve for the airflow generation port. Delay the timing and valve closing timing so that the closing timing of the second intake valve is retarded from the closing timing of the first intake valve, and the opening timing of the second intake valve is the opening timing of the first intake valve The straight port is formed so that the flow coefficient is larger than that of the airflow generation port, and the intake port upstream of the intake port is located inside the intake duct. Open the entire internal flow path The second intake valve has a valve opening timing and a valve closing timing that continuously change so that the required intake air amount is supplied to the combustion chamber, and the required intake air amount decreases. And an operating region in which the closing timing of the second intake valve is continuously retarded while the throttle valve is fully open and the mechanical compression ratio is continuously increased to maintain the actual compression ratio constant. It is characterized by.

  A spark ignition internal combustion engine according to a second aspect of the present invention is the spark ignition internal combustion engine according to the first aspect, wherein the intake of the airflow generation port is upstream of a branch portion between the airflow generation port and the straight port. In the case where a control valve for controlling the flow rate is arranged, the air flow generation port and the straight port are separated by a partition plate between the branch portion and the control valve.

According to the spark ignition internal combustion engine of the first aspect of the present invention, the variable compression ratio mechanism, the air flow generation port, and the straight port are provided, and the intake air amount is controlled by changing the closing timing of the intake valve. In the spark ignition internal combustion engine, when reducing the intake air amount, the opening timing of the second intake valve for the straight port and the opening timing of the first intake valve for the airflow generation port are not changed, Delay the valve closing timing so that the closing timing of the second intake valve is retarded from the closing timing of the first intake valve, and when the throttle valve is fully open, the closing timing of the second intake valve Is continuously retarded and the mechanical compression ratio is continuously increased to keep the actual compression ratio constant. More this arrangement, the intake of the cylinder to the valve closing timing of the second intake valve during the compression stroke is returned to the intake system, it is possible to reduce the intake air amount, also the opening timing of the second intake valve first By setting the intake valve to be retarded from the opening timing, the intake stroke is initially supplied to the cylinder using only the airflow generation port by opening only the first intake valve. Since a strong airflow can be generated, strong turbulence is generated in the cylinder at the ignition timing, and better combustion can be realized compared to the conventional case.

  A spark ignition internal combustion engine according to a second aspect of the present invention is the spark ignition internal combustion engine according to the first aspect, wherein the intake air flow rate of the airflow generation port is controlled upstream of the branch portion between the airflow generation port and the straight port. When the control valve to be installed is arranged, the air flow generating port and the straight port are separated by a partition plate between the branching portion and the control valve, so that the air flow generating port can be satisfactorily controlled by the control valve. The amount of intake air passing therethrough can be controlled.

1 is an overall view of an 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 top sectional view of an internal-combustion engine showing a section of an intake port roughly.

  FIG. 1 is a side sectional view of an internal combustion engine having a variable compression ratio mechanism according to the present invention. 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 catalyst device 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.

  As shown in FIG. 1, a relative position sensor 22 for detecting a relative positional relationship between the crankcase 1 and the cylinder block 2 is attached to the crankcase 1 and the cylinder block 2. An output signal indicating a change in the interval between the crankcase 1 and the cylinder block 2 is output. The variable valve timing mechanism B is provided with a valve timing sensor 23 for generating an output signal indicating the closing timing of the intake valve 7, and an output signal indicating the throttle valve opening is provided to the actuator 16 for driving the throttle valve. A throttle opening sensor 24 is attached.

  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. Output signals of the intake air amount detector 18, the air-fuel ratio sensor 21, the relative position sensor 22, the valve timing sensor 23, and the throttle opening sensor 24 are input to the input port 35 via the corresponding AD converters 37. 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 58 is rotatably inserted into each cam insertion hole 53. It is fixed. These circular cams 58 are coaxial with the rotational axes of the camshafts 54 and 55. On the other hand, on both sides of each circular cam 58, as shown in FIG. 3, an eccentric shaft 57 eccentrically arranged with respect to the rotation axis of each camshaft 54, 55 extends. 56 is mounted eccentrically and rotatable. As shown in FIG. 2, these circular cams 56 are arranged on both sides of each circular cam 58, and these circular cams 56 are rotatably inserted into the corresponding cam insertion holes 51. As shown in FIG. 2, a cam rotation angle sensor 25 that generates an output signal representing the rotation angle of the camshaft 55 is attached to the camshaft 55.

  When the circular cams 58 fixed on the camshafts 54 and 55 are rotated in opposite directions as indicated by arrows in FIG. 3A from the state shown in FIG. 3A, the eccentric shafts 57 are separated from each other. In order to move in the direction, the circular cam 56 rotates in the opposite direction to the circular cam 58 in the cam insertion hole 51, and as shown in FIG. 3B, the position of the eccentric shaft 57 is changed from the high position to the intermediate height position. It becomes. Next, when the circular cam 58 is further rotated in the direction indicated by the arrow, the eccentric shaft 57 is at the lowest position as shown in FIG.

  3A, 3B, and 3C show the positional relationship among the center a of the circular cam 58, the center b of the eccentric shaft 57, and the center c of the circular cam 56 in each state. It is shown.

  3A to 3C, the relative positions of the crankcase 1 and the cylinder block 2 are determined by the distance between the center a of the circular cam 58 and the center c of the circular cam 56. As the distance between the center a of 58 and the center c of the circular cam 56 increases, the cylinder block 2 moves away from the crankcase 1. That is, the variable compression ratio mechanism A changes the relative position between the crankcase 1 and the cylinder block 2 by a crank mechanism using a rotating cam. 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 opposite directions, a pair of worms 61 and 62 having opposite spiral directions are attached to the rotation shaft of the drive motor 59, respectively. Worm wheels 63 and 64 that mesh with the worms 61 and 62 are fixed to the ends 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.

  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 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 FIGS. 1 and 4 is an example. For example, when the intake valve is opened and closed by an electromagnetic or hydraulic actuator, the variable valve timing mechanism can be arbitrarily set by an electromagnetic or hydraulic actuator. In addition, the opening timing and closing timing of the intake valve can be controlled.

  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, the super 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 as well, the combustion chamber volume is set to 50 ml, and the stroke volume of the piston is set to 500 ml, similarly to the example 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 the 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, under such circumstances, it is considered to increase the theoretical thermal efficiency while strictly dividing the mechanical compression ratio and the actual compression ratio. As a result, the theoretical thermal efficiency is governed by the expansion ratio, and the actual compression ratio is compared to the theoretical thermal efficiency. Was 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 ε 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 overall operation control will be schematically described with reference to FIG. FIG. 9 shows changes in the intake air amount, the intake valve closing timing, the mechanical compression ratio, the expansion ratio, the actual compression ratio, and the opening degree of the throttle valve 17 according to the engine load at a certain engine speed. . 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 catalyst device 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 valve closing timing of the intake valve 7 is advanced as shown by the solid line in FIG. ing. At this time, the amount of intake air is large, and at this time, the opening degree of the throttle valve 17 is kept fully open, so that the pumping loss is zero.

  On the other hand, as shown in FIG. 9, when the engine load becomes low, the closing timing of the intake valve 7 is delayed in order 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 reduced to a medium load L1 slightly close to the low load, the mechanical compression ratio becomes a limit mechanical compression ratio (upper limit mechanical compression) that becomes a structural limit of the combustion chamber Ratio). When the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is held at the limit mechanical compression ratio in a region where the load is lower than the engine load L1 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 L1, 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 reduced in a region where the load is lower than the engine load L1 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 L1 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, and the opening degree of the throttle valve 17 is made smaller as the engine load becomes lower.

  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 this embodiment, the variable compression ratio mechanism A is formed so that the expansion ratio is 20 or more.

  As shown in FIG. 10, the spark ignition internal combustion engine of the present embodiment is an intake two-valve type having a pair of intake ports 8a and 8b. In order to realize good combustion with a fast combustion speed by generating turbulence in the cylinder at the ignition timing, a strong air flow such as swirl flow (lateral swirl flow) or tumble flow (vertical swirl flow) in the cylinder during the intake stroke Is preferably generated. Therefore, one intake port 8a is an airflow generation port for generating an airflow in a cylinder such as a swirl port or a tumble port. Since the air flow generation port inevitably has a small flow coefficient, the other intake port 8b is a straight port having a large flow coefficient in order to prevent intake shortage.

  In this internal combustion engine, the intake air amount when the throttle valve 17 is fully opened is not controlled by simultaneously changing the closing timing of both intake valves 7a and 7b, but only the second intake valve 7b for the straight port 8b. As shown in FIG. 9, the valve closing timing is controlled so as to retard as the engine load decreases. That is, changes in the opening timing and closing timing of the intake valve shown in FIG. 5 are applied only to the second intake valve 7b for the straight port 8b, and the camshaft of the first intake valve 7a for the airflow generation port 8a is The variable valve timing mechanism B is not provided, and the valve opening timing and valve closing timing are fixed to the timing shown by the solid line in FIG. 5 when the valve opening period and the lift amount are the same as those of the second intake valve 7b. The In this embodiment, as shown in FIG. 10, the second intake valve 7b for the straight port 8b is made larger than the first intake valve 7a for the airflow generation port 8a.

  Thus, when the intake air amount is maximized, the cam phase of the second intake valve 7b is advanced by the variable valve timing mechanism B, and both the first intake valve 7a and the second intake valve 7b are indicated by solid lines in FIG. The valve opening timing and the valve closing timing are set, and the valves are opened and closed simultaneously.

  On the other hand, when the engine load is reduced and the intake air amount is decreased, the opening timing and closing timing of the second intake valve 7b are retarded, and the closing timing of the second intake valve 7b is set to the first intake valve 7a. In the compression stroke, the intake air in the cylinder is returned to the intake system until the valve closing timing of the second intake valve 7b. At this time, since the opening timing of the second intake valve 7b is set to be retarded from the opening timing of the first intake valve 7a, the air flow is initially caused by opening only the first intake valve 7a in the intake stroke. Since the intake air is supplied into the cylinder using only the generation port 8a and there is no interference with the intake air supplied using the straight port 8b, a strong air flow can be generated in the cylinder. Good turbulence can be generated in the cylinder to achieve good combustion.

  By the way, if the airflow generation port 8a is provided with a control valve CV for controlling the intake flow rate, the engine load L1 that most delays the closing timing of the second intake valve 7b for the straight port 8b (see FIG. 9). When the engine load becomes lower and the intake air amount is further reduced, the control valve CV can be controlled to the valve closing side instead of the throttle valve 17.

  Thereby, since the intake flow rate supplied to the cylinder through the airflow generation port 8a is reduced, the intake air amount in the cylinder can be reduced. In this case, when the control valve CV is controlled to the closed side, the intake air passing through the airflow generation port 8a is easily supplied into the cylinder along the wall surface of the airflow generation port 8a, and a stronger airflow is generated in the cylinder. can do. Thereby, the combustion at this time can be further improved. When it is necessary to reduce the intake air amount further than the intake air amount supplied into the cylinder using both the air flow generation port 8a and the straight port 8b when the control valve CV is closed to the maximum, the throttle valve 17 Is controlled to the valve closing side.

  Such a control valve CV requires an actuator AC. In consideration of such an arrangement of the actuator AC, as shown in FIG. 10, the control valve CV may be arranged upstream of the branch portion BR between the airflow generation port 8a and the straight port 8b. At this time, it is preferable to separate the branch portion BR and the control valve CV into the airflow generation port 8a and the straight port 8b by the partition plate PW. As a result, the intake flow rate that passes through the airflow generation port 8a can be well controlled by the control valve CV.

  The fuel injection valve 13 may be disposed on the upstream side of the control valve CV. However, particularly when the control valve CV is closed, a part of the injected fuel adheres to the control valve CV and is difficult to be supplied into the cylinder. Therefore, it is preferable that the air flow generation port 8a is disposed downstream of the control valve CV as shown by 13a in FIG. 10 or the straight port 8b as shown by 13b in FIG.

  When both the first fuel injection valve 13a of the airflow generation port 8a and the second fuel injection valve 13b of the straight port 8b are provided and intake synchronous injection is performed, the intake air amount is less than the set amount. When the valve opening timing retarded together with the closing timing of the second intake valve 7b for the straight port 8b is retarded from the set timing, all of the required fuel amount is injected by the first fuel injection valve 13a. Therefore, at the beginning of the intake stroke when the second intake valve 7b is not opened, the injected fuel is supplied into the cylinder together with the intake air passing through the air flow generation port 8a, and the air flow generated in the cylinder is injected. Thus, a sufficient homogeneous mixture can be formed in the cylinder by dispersing it throughout the cylinder in a sufficient time until the ignition timing.

  Further, when the intake air amount is equal to or larger than the set amount and the opening timing of the second intake valve 7b for the straight port 8b is the set timing or an advance side from the set timing, the first fuel injection valve 13a and the second fuel It is preferable to inject half of the required fuel amount by the injection valve 13b. At this time, since the required amount of fuel becomes relatively large, only the first fuel injection valve 13a performs fuel injection until the late stage of the intake stroke, and the homogeneous mixing of the fuel injected in the late stage of the intake stroke tends to be insufficient. On the other hand, at this time, the opening of the second intake valve 7b is also relatively quick, and the fuel injected by the second fuel injection valve 13b can also be supplied into the cylinder in the first half of the intake stroke. Fuel injection can be completed by the middle of the intake stroke by both the injection valve 13a and the second fuel injection valve 13b, and a good homogeneous mixture can be formed in the cylinder by the ignition timing.

  The spark ignition internal combustion engine of the present embodiment includes a variable compression ratio mechanism A and a mechanical compression ratio so that the actual compression ratio is constant with respect to the delay of the closing timing of the second intake valve 7b for the straight port 8b. The closing timing of the second intake valve 7b is greatly retarded until the latter half of the compression stroke. Accordingly, if the opening timing of the second intake valve 7b is also relatively retarded, only the first intake valve 7a is opened in the intake stroke, and only the airflow generation port 8a is used to intake air into the cylinder. The period during which the gas is supplied into the cylinder becomes longer, and a strong airflow can be generated in the cylinder. However, the present invention is not limited to an internal combustion engine provided with such a variable compression ratio mechanism A, but can be applied to a spark ignition internal combustion engine that simply controls the intake air amount by changing the closing timing of the intake valve. It is.

DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 7a 1st intake valve 7b 2nd intake valve 8a Airflow generation port 8b Straight port A Variable compression ratio mechanism B Variable valve timing mechanism CV Control valve

Claims (2)

In a spark ignition internal combustion engine that includes a variable compression ratio mechanism and an intake port including an airflow generation port and a straight port , and controls the intake air amount by changing the closing timing of the intake valve.
When reducing the intake air amount, the opening timing and closing timing of the second intake valve for the straight port are set without changing the opening timing and closing timing of the first intake valve for the airflow generation port. The valve closing timing of the second intake valve is retarded so that it is retarded from the valve closing timing of the first intake valve, and the valve opening timing of the second intake valve is set to the valve opening timing of the first intake valve. Implement control to make it more retarded ,
The straight port is formed to have a larger flow coefficient than the airflow generation port,
Inside the intake duct upstream from the intake port, a throttle valve that opens and closes the entire flow path inside the intake duct is disposed,
The second intake valve continuously changes its valve opening timing and valve closing timing so as to supply the required amount of intake air to the combustion chamber,
When the required intake air amount decreases, with the throttle valve fully open, the closing timing of the second intake valve is continuously retarded and the mechanical compression ratio is continuously increased to increase the actual compression ratio. and having an operating area to maintain constant, spark ignition internal combustion engine. In the case where a control valve for controlling the intake air flow rate of the airflow generation port is disposed upstream of the branch portion between the airflow generation port and the straight port, a partition plate is provided between the branch portion and the control valve. separating the said air flow generating port and the straight port, spark ignition internal combustion engine according to claim 1.

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