JP4631848B2 - Spark ignition internal combustion engine - Google Patents

Description

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

  In a normal internal combustion engine, when the exhaust valve is opened, the pressure in the cylinder is reduced to the pressure in the exhaust port during the exhaust stroke, for example, atmospheric pressure. However, if the expansion ratio is increased, the pressure in the cylinder may become atmospheric pressure or lower in the latter half of the expansion stroke, and the pressure in the cylinder may increase when the exhaust valve is opened. However, if the pressure in the cylinder falls below atmospheric pressure in the latter half of the expansion stroke, a pumping loss occurs.

Therefore, in order to prevent the occurrence of such a pumping loss, an internal combustion engine provided with an air valve for opening the supercharged air into the cylinder only in the latter half of the expansion stroke in addition to the normal intake valve. It is known (see Patent Document 1). In this internal combustion engine, the pressure in the cylinder in the latter half of the expansion stroke is increased by the air supplied from the air valve, and thus the pressure in the cylinder is prevented from becoming below atmospheric pressure.
JP 2001-152890 A

  However, when the expansion ratio is increased, the exhaust gas temperature is lowered. At this time, when air is supplied from the air valve into the cylinder in the latter half of the expansion stroke, the exhaust gas temperature is further lowered. As a result, there is a problem that the temperature of the catalyst disposed in the exhaust passage is lowered, so that the catalyst becomes inactive and thus the exhaust purification action is impaired. Furthermore, when such an air valve is arranged in addition to a normal intake valve, there is a problem that the structure becomes complicated.

In order to solve the above problem, according to the present invention, a variable valve timing mechanism capable of controlling the opening timing of the exhaust valve is provided, and when the exhaust valve is opened, the pressure in the cylinder is equal to or lower than the atmospheric pressure. When the exhaust valve opens, the opening timing of the exhaust valve is kept away from the exhaust bottom dead center.If the pressure in the cylinder is higher than the atmospheric pressure when the exhaust valve opens, the opening timing of the exhaust valve becomes the exhaust bottom dead center. To be close to.

  The exhaust gas temperature can be kept high while preventing the occurrence of pumping loss.

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 the outlet of the compressor 15a of the exhaust turbocharger 15 via an intake duct 14, and the inlet of the compressor 15a is connected to an air cleaner 17 via an intake air amount detector 16 using, for example, heat rays. A throttle valve 19 driven by an actuator 18 is disposed in the intake duct 14.

  On the other hand, the exhaust port 10 is connected to an inlet of an exhaust turbine 15b of an exhaust turbocharger 15 via an exhaust manifold 20, and an outlet of the exhaust turbine 15b is connected to a catalytic converter 22 containing, for example, a three-way catalyst via an exhaust pipe 21. Is done. An air-fuel ratio sensor 23 is disposed in the exhaust pipe 21.

  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 when performing the operation, and the valve closing timing of the intake valve 7 can be controlled and the intake air can be controlled in order to change the actual start timing of the compression action. A variable valve timing mechanism B that can individually control the opening timing of the valve 7 and a variable valve timing mechanism C that can control the opening timing of the exhaust valve 9 and can also control the closing timing of the exhaust valve 9 individually. And are provided.

  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 16 and the output signal of the air-fuel ratio sensor 23 are input to the input port 35 via the corresponding AD converter 37. A pressure sensor 39 for detecting the pressure in the cylinder is disposed in the combustion chamber 5, and an output signal of the pressure sensor 39 is input to the input port 35 via the corresponding AD converter 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 connected to the input port 35 via the corresponding AD converter 37. Entered. Further, the input port 35 is connected to a crank angle sensor 42 that generates an output pulse every time the crankshaft rotates, for example, 30 °. On the other hand, the output port 36 is connected to the spark plug 6, the fuel injection valve 13, the throttle valve drive actuator 18, the variable compression ratio mechanism A ′, the variable valve timing mechanisms B and C via the corresponding drive circuit 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 opposite directions, a pair of worm gears 61 and 62 having screw directions opposite to each other are attached to the rotation shaft of the drive motor 59, respectively. Gears 63 and 64 that mesh 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.

  4 shows a variable valve timing mechanism B provided for the camshaft 70 for driving the intake valve 7 in FIG. As shown in FIG. 4, the variable valve timing mechanism B is attached to one end of the camshaft 70 to change the cam phase of the camshaft 70, and the camshaft 70 and intake valve 7 valve lifters. The cam operating angle changing portion B2 is disposed between the cam shafts 70 and changes the cam operating angle of the camshaft 70 to a different operating angle and transmits the cam operating angle to the intake valve 7. FIG. 4 shows a side sectional view and a plan view of the cam working angle changing portion B2.

  First, the cam phase changing portion B1 of the variable valve timing mechanism B will be described. The cam phase changing portion B1 is rotated by a crankshaft of the engine via a timing belt in a direction indicated by an arrow, a timing pulley 71, A cylindrical housing 72 that rotates together, a rotating shaft 73 that rotates together with the camshaft 70 and that can rotate relative to the cylindrical housing 72, and an outer peripheral surface of the rotating shaft 73 from an inner peripheral surface of the cylindrical housing 72 And a plurality of partition walls 74 extending between the partition walls 74 and the outer peripheral surface of the rotating shaft 73 to the inner peripheral surface of the cylindrical housing 72. An advance hydraulic chamber 76 and a retard hydraulic chamber 77 are formed respectively.

  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 camshaft 70 is to be advanced, the spool valve 85 is moved downward in FIG. 4, and the hydraulic oil supplied from the supply port 82 enters the advance angle hydraulic chamber 76 via the hydraulic port 79. The hydraulic oil in the retarding hydraulic chamber 77 is supplied and discharged from the drain port 84. At this time, the rotation shaft 73 is rotated relative to the cylindrical housing 72 in the arrow X direction.

  On the other hand, when the cam phase of the camshaft 70 is to be retarded, the spool valve 85 is moved upward in FIG. 4 and the hydraulic oil supplied from the supply port 82 is used for retarding via the hydraulic port 80. The hydraulic oil in the advance hydraulic chamber 76 is discharged from the drain port 83 while being supplied to the 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 X.

  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 cam phase changing portion B1 can advance or retard the cam phase of the camshaft 70 by a desired amount. That is, the valve opening timing of the intake valve 7 can be arbitrarily advanced or retarded by the cam phase changing unit B1.

  Next, the cam working angle changing portion B2 of the variable valve timing mechanism B will be described. The cam working angle changing portion B2 is arranged in parallel with the camshaft 70 and is moved in the axial direction by the actuator 91; An intermediate cam 94 engaged with the cam 92 of the camshaft 70 and slidably fitted in an axially extending spline 93 formed on the control rod 90 and a valve lifter for driving the intake valve 7 And a swing cam 96 slidably fitted to a spline 95 extending in a threaded manner formed on the control rod 90, and a cam 97 is provided on the swing cam 96. Is formed.

When the camshaft 70 rotates, the intermediate cam 94 is always swung by a certain angle by the cam 92, and at this time, the rocking cam 96 is also swung by a certain angle. On the other hand, the intermediate cam 94 and the swing cam 96 are supported so as not to move in the axial direction of the control rod 90. Therefore, when the control rod 90 is moved in the axial direction by the actuator 91, the swing cam 96 is intermediate. It is rotated relative to the cam 94.

When the cam 97 of the swing cam 9 6 when the cam 92 of the camshaft 70 is started engaged with the intermediate cam 94 due to the relative rotational positional relationship between the intermediate cam 94 and swing cam 96 starts engage the valve lifter 24 As shown by a in FIG. 5B, the valve opening period and the lift of the intake valve 7 become the largest. On the other hand, when the swing cam 96 is rotated relative to the intermediate cam 94 in the direction of the arrow Y by the actuator 91, the cam 92 of the camshaft 70 is engaged with the intermediate cam 94 for a while. Thereafter, the cam 97 of the swing cam 96 is engaged with the valve lifter 24. In this case, as indicated by b in FIG. 5B, the valve opening period and the lift amount of the intake valve 7 are smaller than a.

  When the swing cam 96 is further rotated relative to the intermediate cam 94 in the direction of the arrow Y in FIG. 4, the valve opening period and the lift amount of the intake valve 7 are further reduced as indicated by c in FIG. . That is, the valve opening period of the intake valve 7 can be arbitrarily changed by changing the relative rotational position of the intermediate cam 94 and the swing cam 96 by the actuator 91. However, in this case, the lift amount of the intake valve 7 becomes smaller as the opening period of the intake valve 7 becomes shorter.

  Thus, the cam phase changing unit B1 can arbitrarily change the valve opening timing of the intake valve 7, and the cam operating angle changing unit B2 can arbitrarily change the valve opening period of the intake valve 7, so that the cam phase By both the change part B1 and the cam working angle change part B2, that is, by the variable valve timing mechanism B, the valve opening timing and the valve opening period of the intake valve 7, that is, the valve opening timing and the valve closing timing of the intake valve 7 are set. It can be changed arbitrarily.

  On the other hand, the variable valve timing mechanism C provided for the camshaft 69 for driving the exhaust valve 9 in FIG. 1 is the variable valve timing provided for the camshaft 70 for driving the intake valve 7. Therefore, the variable valve timing mechanism C can arbitrarily change the opening timing and closing timing of the exhaust valve 9. In the embodiment according to the present invention, as shown by a, b, c and d in FIG. 6, the closing timing of the exhaust valve 9 is fixed, and the opening timing of the exhaust valve 9 is controlled.

  The variable valve timing mechanism B shown in FIGS. 1 and 4 shows an example, and various types of variable valve timing mechanisms other than the examples shown in FIGS. 1 and 4 can be used.

  Next, the meanings of terms used in the present application will be described with reference to FIG. 7A, 7B and 7C 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. 7A 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. 7A, this mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11.

  FIG. 7B explains 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. 7B, 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 after 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. 7B, the actual compression ratio is (50 ml + 450 ml) / 50 ml = 10.

  FIG. 7C 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. 7C, 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. 8 shows the relationship between the theoretical thermal efficiency and the expansion ratio, and FIG. 9 shows a comparison between a normal cycle and an ultra-high expansion ratio cycle that are selectively used according to the load in the present invention.

  FIG. 9A 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 compression bottom dead center. In the example shown in FIG. 9A, the combustion chamber volume is 50 ml, and the stroke volume of the piston is 500 ml, as in the examples shown in FIGS. 7A, 7B, and 7C. As can be seen from FIG. 9A, in the 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. 8 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, 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, the actual compression ratio should be increased. However, the actual compression ratio can only be increased to a maximum of about 12 due to the restriction of the occurrence of knocking during 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 in FIG. 8 shows 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 maintaining the actual compression ratio 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. 9B 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. 9B, 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. 9A, the actual compression ratio is almost 11 and the expansion ratio is 11 as described above. Compared to this case, only the expansion ratio is shown in FIG. 9B. It can be seen that it has been increased to 26. This is why it is called an ultra-high expansion ratio cycle.

  As described above, in general, in an internal combustion engine, the lower the engine load, the lower the thermal efficiency. Therefore, in order to improve the thermal efficiency when the vehicle is running, that is, to improve the fuel efficiency, the thermal efficiency during the engine low load operation is reduced. It is necessary to improve. On the other hand, in the ultra-high expansion ratio cycle shown in FIG. 9B, 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, the engine is operated at a very high expansion ratio cycle as shown in FIG. 9B during low engine load operation, and the normal cycle as shown in FIG. 9A during engine high load operation.

FIG. 10 shows the overall operation control during steady operation at a low engine speed. Hereinafter, the overall operation control will be described with reference to FIG.
FIG. 10 shows changes in mechanical compression ratio, expansion ratio, intake valve 7 closing timing, actual compression ratio, intake air amount, throttle valve 19 opening degree, and pumping loss according to engine load. . In the embodiment according to the present invention, the average air-fuel ratio in the normal combustion chamber 5 is the air-fuel ratio sensor 23 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 22. Feedback control to the stoichiometric air-fuel ratio based on the output signal.

As described above, the normal cycle shown in FIG. 9A is executed during engine high load operation. Accordingly, as shown in FIG. 10, the mechanical compression ratio is lowered at this time, so the expansion ratio is low, and the closing timing of the intake valve 7 is advanced as shown by the solid line in FIG. Further, at this time, the intake air amount is large, and at this time, the opening degree of the throttle valve 19 is kept fully open or almost fully open, so that the pumping loss is zero.

On the other hand, as shown in FIG. 10, when the engine load is lowered, the mechanical compression ratio is increased accordingly, and thus the expansion ratio is also increased. Further, at this time, the valve closing timing of the intake valve 7 is delayed as the engine load becomes lower as shown by the solid line in FIG. 10 so that the actual compression ratio is kept substantially constant. At this time as well, the throttle valve 19 is kept fully open or substantially fully open, so that the amount of intake air supplied into the combustion chamber 5 does not depend on the throttle valve 19 but the closing timing of the intake valve 7. Is controlled by changing. Also at this time, the pumping loss becomes zero.

  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, since the air-fuel ratio in the combustion chamber 5 is the stoichiometric air-fuel ratio, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the fuel amount. Become.

When the engine load is further reduced, the mechanical compression ratio is further increased. When the mechanical compression ratio reaches the limit mechanical compression ratio that is the structural limit of the combustion chamber 5, the engine when the mechanical compression ratio reaches the limit mechanical compression ratio is reached. In the region where the load is lower than the load L ₁ , the mechanical compression ratio is maintained at the limit mechanical compression ratio. Therefore, the mechanical compression ratio is maximized and the expansion ratio is maximized during engine low load operation. In other words, in the embodiment according to the present invention, the mechanical compression ratio is maximized so that the maximum expansion ratio is obtained during engine low load operation. Further, at this time, the actual compression ratio is maintained at substantially the same actual compression ratio as that during the engine medium and high load operation.

On the other hand, as shown by the solid line in FIG. 10, the closing timing of the intake valve 7 is delayed to the limit closing timing that can control the amount of intake air supplied into the combustion chamber 5 as the engine load becomes lower. closing timing of the valve 7 is the closing timing of the intake valve 7 in the region of a load lower than the engine load L ₂ when reaching the limit closing timing is held at the limit closing timing. If the closing timing of the intake valve 7 is held at the limit closing timing, the intake air amount can no longer be controlled by the change in the closing timing of the intake valve 7, so it is necessary to control the intake air amount by some other method There is.

In the embodiment shown in FIG. 10, at this time, that is, in a region where the load is lower than the engine load L ₂ when the closing timing of the intake valve 7 reaches the limit closing timing, the throttle valve 19 enters the combustion chamber 5. The amount of intake air supplied is controlled. However, when the intake air amount is controlled by the throttle valve 19 , the pumping loss increases as shown in FIG.

In order to prevent such a pumping loss, the throttle valve 19 is fully opened or substantially fully opened in a region where the load is lower than the engine load L ₂ when the closing timing of the intake valve 7 reaches the limit closing timing. The air-fuel ratio can be increased as the engine load becomes lower in the held state. At this time, it is preferable to arrange the fuel injection valve 13 in the combustion chamber 5 and perform stratified combustion.

  As shown in FIG. 10, the actual compression ratio is maintained substantially constant regardless of the engine load at the time of low engine speed. However, if the engine speed increases, the air-fuel mixture in the combustion chamber 5 is disturbed, so that knocking does not easily occur. Therefore, in the embodiment according to the present invention, the actual compression ratio increases as the engine speed increases.

  On the other hand, as described above, the expansion ratio is 26 in the ultra-high expansion ratio cycle shown in FIG. This expansion ratio is preferably as high as possible, but if it is 20 or more, a considerably high theoretical thermal efficiency can be obtained. Therefore, in the present invention, the variable compression ratio mechanism A is formed so that the expansion ratio is 20 or more. In the example shown in FIG. 9, the mechanical compression ratio is continuously changed according to the engine load. However, the mechanical compression ratio can be changed stepwise according to the engine load.

On the other hand, the intake air amount can be controlled without depending on the throttle valve 19 by advancing the closing timing of the intake valve 7 as the engine load becomes lower as shown by the broken line in FIG. Therefore, if expressed in such a way as to include both the case indicated by the solid line and the case indicated by the broken line in FIG. 10, in the embodiment according to the present invention, the closing timing of the intake valve 7 increases as the engine load decreases. until the limit closing timing L ₂ enabling control of the amount of intake air fed into the combustion chamber will be kept away from the intake bottom dead center BDC.

  Next, the control of the opening timing of the exhaust valve 9 according to the present invention will be described with reference to FIGS. In FIGS. 11 to 14, (A) shows the valve opening periods of the intake valve 7 and the exhaust valve 9. In FIGS. 11 to 14 (A), the IC indicates the closing timing of the intake valve 7. EO indicates the opening timing of the exhaust valve 9. On the other hand, (B) in FIG. 11 to FIG. 14 is a PV diagram showing the relationship between the pressure P in the combustion chamber 5, that is, the pressure in the cylinder, and the volume V in the cylinder in logarithm.

  FIG. 11 shows a case where the operation is performed in a normal cycle. At this time, as shown in (B), when the exhaust valve 9 is opened in EO, the pressure P in the cylinder decreases to the atmospheric pressure PX, strictly speaking, to almost the atmospheric pressure, and then during the exhaust stroke, The internal pressure P is maintained at the atmospheric pressure PX. At this time, the throttle valve 19 is fully opened or substantially fully opened, and the amount of intake air trapped in the cylinder is controlled by changing the valve closing timing IC of the intake valve 7.

  On the other hand, FIG. 12 shows the expansion ratio until the pressure P in the cylinder decreases to below atmospheric pressure in the latter half of the expansion stroke, and thus the pressure P in the cylinder increases when the exhaust valve 9 is opened. Shows the case where is increased. In this case, as shown in FIG. 12, the closing timing IC of the intake valve 7 is considerably retarded, and the mechanical compression ratio is made considerably high so that the actual compression ratio is not lowered. The amount of decrease in the pressure P in the cylinder until the expansion stroke is completed is proportional to the expansion ratio. Therefore, when the expansion ratio increases, the pressure P in the cylinder becomes equal to or lower than the atmospheric pressure as shown by hatching in FIG. A condition occurs and thus a pumping loss occurs.

  Therefore, in the present invention, in order to prevent the occurrence of such a pumping loss, as shown in FIG. 13 or FIG. 14, the exhaust valve 9 prevents the pressure P in the cylinder from increasing when the exhaust valve 9 is opened. The valve opening timing EO is kept away from the exhaust bottom dead center. In this case, in the embodiment shown in FIG. 13, the opening timing EO of the exhaust valve 9 is moved away from the exhaust bottom dead center by advancing the opening timing EO of the exhaust valve 9, and the implementation shown in FIG. In the example, the valve opening timing EO of the exhaust valve 9 is moved away from the exhaust bottom dead center by retarding the valve opening timing EO of the exhaust valve 9.

  Specifically, in the embodiment shown in FIG. 13, the exhaust valve 9 is opened when the pressure P in the cylinder drops to the atmospheric pressure PX. As a result, as shown in FIG. 13B, the pressure P in the cylinder does not become equal to or lower than the atmospheric pressure PX, and thus it is possible to prevent a pumping loss from occurring.

  On the other hand, in the embodiment shown in FIG. 14, the exhaust valve 9 is opened when the pressure P in the cylinder becomes equal to or lower than the atmospheric pressure PX, and then the piston 4 rises and the pressure P in the cylinder rises to the atmospheric pressure PX. It is done. In this case, no pumping loss occurs. In this case, the exhaust gas once exhausted into the exhaust port 10 and cooled is not returned into the cylinder, so that deposits are prevented from accumulating on the top surfaces of the exhaust valve 9 and the piston 4. Can do.

  FIG. 15 shows a control routine for the valve opening timing EO when the valve opening timing EO of the exhaust valve 9 is controlled so as not to cause a pumping loss when operating with an ultra-high expansion ratio cycle. . This control routine is executed by interruption every predetermined time. In this control routine, the valve opening timing EO of the exhaust valve 9 is always before the intake top dead center.

Referring to FIG. 15, first, at step 100, it is determined whether or not the operation is being performed with an ultra-high expansion ratio cycle. When the operation is not performed with the super high expansion ratio cycle, the routine proceeds to step 101, where the valve opening timing EO of the exhaust valve 9 is made the normal valve opening timing EO ₀ shown in FIG. On the other hand, when the operation is performed with the ultra-high expansion ratio cycle, the routine proceeds to step 102 where it is determined whether or not it is the second half of the expansion stroke, and when the second half of the expansion stroke, the routine proceeds to step 103 and the pressure sensor 39 is reached. Thus, the pressure P in the cylinder is detected.

Next, at step 104, it is judged if the pressure P in the cylinder has dropped below the atmospheric pressure PX. When P <PX, the routine proceeds to step 105, where the constant value α is added to the correction advance amount ΔEO. Next, in step 106, the opening timing EO of the exhaust valve 9 is calculated by adding the advance correction amount ΔEO normal opening timing EO _0. When the valve opening timing EO of the exhaust valve 9 is calculated, the valve opening timing of the exhaust valve 9 is controlled by the variable valve timing mechanism C to the calculated valve opening timing EO. Thus, when P <PX, the valve opening timing EO of the exhaust valve 9 is advanced until the pressure P in the cylinder reaches the atmospheric pressure PX.

  On the other hand, when it is determined in step 104 that P ≧ PX, the routine proceeds to step 107 where the pressure P in the cylinder immediately before the exhaust valve 9 is opened is a value obtained by adding a small constant value ΔP to the atmospheric pressure PX (PX + ΔP). Is also determined. When the pressure P in the cylinder immediately before the opening of the exhaust valve 9 is higher than the atmospheric pressure PX by ΔP or more, the routine proceeds to step 108 where the constant value β is subtracted from the corrected advance value ΔEO, and then the routine proceeds to step 106. That is, when the pressure P in the cylinder immediately before opening the exhaust valve 9 is higher than the atmospheric pressure PX by ΔP or more, the valve opening timing EO of the exhaust valve 9 is retarded.

  In general, the control of the valve opening timing EO of the exhaust valve 9 is expressed by the exhaust valve 9 when the expansion ratio is increased until the pressure in the cylinder increases when the exhaust valve 9 is opened. The valve opening timing EO of the exhaust valve 9 is kept away from the exhaust bottom dead center until the pressure P in the cylinder does not increase when the valve is opened, and the pressure P in the cylinder is predetermined when the exhaust valve 9 is opened. When the pressure ΔP is reduced by more than the pressure ΔP, the opening timing EO of the exhaust valve 9 is brought close to the exhaust bottom dead center.

  As can be seen from the above description, in the embodiment shown in FIG. 15, the pressure P in the cylinder immediately before the exhaust valve 9 is opened is from the atmospheric pressure PX to a pressure (PX + ΔP) higher than the atmospheric pressure PX by a certain value ΔP. Maintained during. Therefore, the occurrence of pumping loss can be prevented, and the exhaust valve 9 does not open until the pressure P in the cylinder sufficiently decreases during the expansion stroke, so that the thermal efficiency can be improved.

  In the embodiment shown in FIG. 15, the pressure P in the cylinder is detected by the pressure sensor 39, but the pressure P in the cylinder is estimated from the engine speed, engine load, fuel supply amount, actual compression ratio, ignition timing, and the like. You can also. In this case, the pressure in the cylinder is calculated from the engine speed, the engine load, the fuel supply amount, the actual compression ratio, the ignition timing, etc. using a model stored in advance. In step 103 in FIG. Instead of using and detecting the pressure P, the pressure in the cylinder calculated from the model is read.

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. It is a figure which shows the lift amount of 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 which shows the valve opening period of an intake valve and an exhaust valve, and the pressure change in a cylinder. It is a figure which shows the valve opening period of an intake valve and an exhaust valve, and the pressure change in a cylinder. It is a figure which shows the valve opening period of an intake valve and an exhaust valve, and the pressure change in a cylinder. It is a figure which shows the valve opening period of an intake valve and an exhaust valve, and the pressure change in a cylinder. It is a flowchart for controlling the valve opening timing of an exhaust valve.

Explanation of symbols

1 Crankcase 2 Cylinder block 3 Cylinder head 4 Piston 5 Combustion chamber 7 Intake valve 9 Exhaust valve A Variable compression ratio mechanism B, C Variable valve timing mechanism

Claims (3)

Equipped with a variable valve timing mechanism that can control the opening timing of the exhaust valve. If the pressure in the cylinder falls below atmospheric pressure when the exhaust valve opens , the opening timing of the exhaust valve is exhausted. A spark ignition type internal combustion engine that is kept away from bottom dead center and that the opening timing of the exhaust valve is close to exhaust bottom dead center when the pressure in the cylinder is equal to or higher than atmospheric pressure when the exhaust valve is opened . The opening timing of the exhaust valve is brought close to the exhaust bottom dead center when the pressure in the cylinder is equal to or higher than the atmospheric pressure when the exhaust valve is opened and the pressure in the cylinder decreases by a predetermined pressure or more. 2. The spark ignition internal combustion engine according to 1. Equipped with a variable compression ratio mechanism that can change the mechanical compression ratio and a variable valve timing mechanism that can control the closing timing of the intake valve. The spark ignition internal combustion engine according to claim 1 , wherein the mechanical compression ratio is increased while being kept away from the point .

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