JP2011122512A - Spark ignition type internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a spark ignition type internal combustion engine in which an intake air amount supplied to a combustion chamber is mainly controlled by changing a closing timing of an intake valve, and which is capable of preventing thermal efficiency from lowering and fuel efficiency from deteriorating which are generated in association with excessive delay of closing timing of the intake valve. <P>SOLUTION: The spark ignition type internal combustion engine includes: a variable compression ratio mechanism A capable of changing a mechanical compression ratio; a variable valve timing mechanism B capable of controlling the closing timing of an intake valve 7; and an intake gas temperature detection means detecting or estimating a temperature of an intake gas supplied to the combustion chamber when the intake valve is closed. During a low load operation of the engine, if a temperature T detected or estimated by the intake gas temperature detection means is higher than a reference temperature Tr, the closing timing of the intake valve is advanced such that the closing timing approaches a compression bottom dead center and the mechanical compression ratio is reduced, compared to a condition in which the temperature T is not higher than the reference temperature Tr. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

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

  It has 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, and the amount of intake air supplied into the combustion chamber is mainly the closing timing of the intake valve There is known a spark ignition type internal combustion engine which is controlled by changing the engine pressure and has a higher mechanical compression ratio at the time of engine low load operation than at the time of engine high load operation (for example, Patent Document 1).

  In particular, since the mechanical compression ratio (that is, the expansion ratio) has a larger influence on the theoretical thermal efficiency than the actual compression ratio, the spark ignition type internal combustion engine described in Patent Document 1 is actually compressed during engine low load operation. While maintaining the ratio low, the mechanical compression ratio is set to a high value of, for example, 20 or more. Thereby, in the spark ignition internal combustion engine described in Patent Document 1, the theoretical thermal efficiency can be made extremely high, and the fuel efficiency is greatly improved accordingly.

JP 2007-303423 A

  By the way, in the spark ignition internal combustion engine disclosed in Patent Document 1, the intake air amount supplied into the combustion chamber is controlled mainly by changing the valve closing timing of the intake valve. Specifically, when the intake air amount supplied into the combustion chamber should be increased, the valve closing timing of the intake valve is advanced so as to approach the compression bottom dead center to reduce the intake air amount supplied into the combustion chamber. When it should be, the closing timing of the intake valve is retarded so as to be away from the compression bottom dead center.

  Here, when the closing timing of the intake valve is retarded so as to be away from the compression bottom dead center, the intake valve is opened while the piston is descending, and then the piston is also raised to some extent. During this period, the intake valve remains open. If the intake valve remains open while the piston is moving up, a part of the intake gas once flowing into the combustion chamber during the lowering of the piston is returned to the intake port as the piston moves up. Will be. The amount of intake gas returned into the intake port increases as the intake valve closing timing is retarded.

  On the other hand, the intake gas once flowing into the combustion chamber is heated by receiving heat from the cylinder wall surface in the combustion chamber. Therefore, the temperature of the intake gas that once flows into the combustion chamber and then returns to the intake port is higher than that before flowing into the combustion chamber. For this reason, the temperature of the intake gas in the intake port increases, and as a result, the temperature of the intake gas flowing into the combustion chamber in the next cycle increases.

  When the temperature of the intake gas flowing into the combustion chamber increases in this way, the temperature of the intake gas in the combustion chamber at the compression top dead center increases, and knocking is likely to occur. For this reason, when the temperature of the intake gas flowing into the combustion chamber rises, it is necessary to retard the ignition timing in order to suppress the occurrence of knocking, resulting in a decrease in thermal efficiency and a deterioration in fuel consumption. I will.

  Accordingly, an object of the present invention is to provide an excessive amount of intake valve closing timing in a spark ignition type internal combustion engine that mainly controls the amount of intake air supplied into the combustion chamber by changing the closing timing of the intake valve. The purpose is to suppress a decrease in thermal efficiency and a deterioration in fuel consumption due to retardation.

  In order to solve the above problems, in the first 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 combustion when the intake valve is closed An intake gas temperature detecting means for detecting or estimating the temperature of the intake gas supplied into the room, and the temperature detected or estimated by the intake gas temperature detecting means is higher than the reference temperature during engine low load operation Sometimes, there is provided a spark ignition internal combustion engine in which the closing timing of the intake valve is advanced so as to approach the compression bottom dead center and the mechanical compression ratio is made smaller than when the temperature is lower than the reference temperature.

  According to a second invention, in the first invention, in the engine low load operation, when the temperature detected or estimated by the intake gas temperature detecting means is higher than the reference temperature, the throttle valve is lower than when the temperature is lower than the reference temperature. The opening of was made small.

  In the third invention, in the first or second invention, the actual compression ratio is made substantially the same regardless of the temperature detected or estimated by the intake gas temperature detecting means at the time of engine low load operation.

  According to a fourth invention, in any one of the first to third inventions, when the temperature detected or estimated by the intake gas temperature detection means is higher than the reference temperature during engine low load operation, the detection or estimation is performed. As the temperature increases, the closing timing of the intake valve is advanced so as to approach the compression bottom dead center, the mechanical compression ratio is decreased, and the opening of the throttle valve is decreased.

  In the fifth invention, in any one of the first to fourth inventions, when the temperature detected or estimated by the intake gas temperature detecting means is equal to or lower than the reference temperature during the engine low load operation, the detected or estimated is detected. Regardless of the temperature, the valve closing timing and mechanical compression ratio of the intake valve are maintained almost constant.

  According to a sixth aspect of the invention, in any one of the first to fifth aspects of the invention, when the internal combustion engine is cold started, the temperature detected or estimated by the intake gas temperature detecting means is assumed to be lower than the reference temperature. The valve closing timing and mechanical compression ratio were controlled.

  According to the present invention, in a spark ignition type internal combustion engine that controls the amount of intake air supplied into the combustion chamber mainly by changing the closing timing of the intake valve, an excessive delay of the closing timing of the intake valve is performed. It is possible to suppress a decrease in thermal efficiency and a deterioration in fuel consumption associated with.

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 flow of the intake gas when the valve closing timing of an intake valve is retarded. It is a figure which shows the relationship between intake temperature and the valve closing timing of an intake valve, etc. 5 is a flowchart showing a control routine for control of intake valve closing timing and the like. It is a figure which shows the map used for calculation of the valve closing timing correction amount etc. of an intake valve. FIG. 12 is a view similar to FIG. 11 showing the relationship between the intake air temperature and the closing timing of the intake valve. FIG. 10 is a view similar to FIG. 9 showing changes in the mechanical compression ratio and the like according to the engine load.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. In the following description, the same reference numerals are assigned to similar components.

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 an 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 when performing the operation, and a variable valve timing mechanism B 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 the corresponding AD converters 37, respectively. A load sensor 41 that generates an output voltage proportional to the amount of depression 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. The 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, 10 °. 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 projecting portions 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 projecting portion 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. A cam insertion hole 53 having a circular cross section is 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 that 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 shown by solid arrows in FIG. 3A from the state shown in FIG. 3 moves toward the lower center, so that 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. As shown, 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 located 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 opposite spiral directions 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 end portions of the camshafts 54 and 55, respectively. In the present 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, the cam phase changer B <b> 1, and the valve lifter of the camshaft 70 and the intake valve 7. The cam operating angle changing portion B2 is arranged between the cam operating angle change portion 26 and the cam operating angle changing portion B2 for changing the operating angle of the cam of the camshaft 70 to a different operating angle. As for the cam working angle changing portion B2, a side sectional view and a plan view are shown in FIG.

  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 rotary 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 rotary 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 rotary 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.

  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 that performs communication cutoff control 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 The relative rotation position at that time is held. Therefore, the cam phase changer B1 can advance or retard the cam phase of the camshaft 70 by a desired amount as shown in FIG. 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 26 and a swing cam 96 slidably fitted to a spirally extending spline 95 formed on the control rod 90, and a cam 97 is formed on the swing cam 96. Has been.

  When the camshaft 70 is rotated, the intermediate cam 94 is always swung by a certain angle by the cam 92. 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 96 starts to engage with the valve lifter 26 when the cam 92 of the camshaft 70 starts to engage with the intermediate cam 94 due to the relative rotational positional relationship between the intermediate cam 94 and the swing cam 96. As shown by a in FIG. 5B, the valve opening period and 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 26. 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 (working angle) 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 valve opening period of the intake valve 7, that is, the valve opening timing and valve closing timing of the intake valve 7 are set. It can be changed arbitrarily.

  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. In particular, in the embodiment of the present invention, any type of mechanism may be used as long as the valve closing timing mechanism can change the valve closing timing of the intake valve 7. A variable valve timing mechanism similar to the variable valve timing mechanism B of the intake valve 7 may be provided for the exhaust valve 9 as well.

  Next, the meanings of terms used in the present application will be described with reference to FIG. For the sake of explanation, FIG. 6 shows an engine having a combustion chamber volume of 50 ml and a piston stroke volume of 500 ml. In FIG. 6, the combustion chamber volume is the value when the piston is located at the compression top dead center. It represents the volume of the combustion chamber.

  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 action is not performed while the intake valve is open, and the actual compression starts from the time when the intake valve is closed. The action begins. Therefore, the actual compression ratio is expressed as 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 most basic control 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 compression 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, as in the example shown in FIG. 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, 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 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, when the theoretical thermal efficiency is increased by strictly dividing the mechanical compression ratio and the actual compression ratio, the expansion ratio dominates the theoretical thermal efficiency, and the actual compression ratio has little influence on the theoretical thermal efficiency. 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. The theoretical thermal efficiency when the expansion ratio is increased while maintaining the actual compression ratio at a low value as described above, and the theory 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 increase in thermal efficiency.

  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 greatly increase. 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. For this reason, the cycle shown in FIG. 8B is referred to as an ultra-high expansion ratio cycle.

  As described above, in general, in an internal combustion engine, the lower the engine load, the worse 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 required. It is necessary to improve. 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, the engine is operated at a very high expansion ratio cycle as shown in FIG. 8B during low engine load operation, and at the normal cycle shown in FIG. 8A during engine high load operation.

Next, the overall operation control will be described with reference to FIG.
FIG. 9 shows changes in the closing timing of the intake valve 7, the mechanical compression ratio, the expansion ratio, the actual compression ratio, the opening degree of the throttle valve 17 and the intake air amount in accordance with the engine load at a certain engine speed. ing. In particular, broken lines in FIG. 9 indicate changes in parameters when basic control according to the present invention (hereinafter referred to as “basic control”) is performed.

In the illustrated example, unburned hydrocarbons (unburned HC), carbon monoxide (CO), and nitrogen oxides (NO _x ) in the exhaust gas can be simultaneously reduced by the three-way catalyst in the catalytic converter 23. The average air-fuel ratio in the normal combustion chamber 5 is feedback-controlled to the theoretical air-fuel ratio based on the output signal from the air-fuel ratio sensor 27.

  Below, the basic control of the spark ignition type internal combustion engine shown by the broken line in FIG. 9 will be described. In the spark ignition type internal combustion engine in the present embodiment, the normal cycle shown in FIG. 8A is executed during engine high load operation as described above. Therefore, as shown by the broken line in FIG. 9, the mechanical compression ratio is lowered during the engine high load operation, the expansion ratio is low, and the closing timing of the intake valve 7 is advanced. 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 or almost fully open. Since the opening degree of the throttle valve 17 is kept fully open or almost fully open in this way, the pumping loss is almost zero.

  On the other hand, as shown by the broken line 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, the mechanical compression ratio is increased as the engine load is lowered as shown by the broken line in FIG. 9 so that the actual compression ratio is kept substantially constant, and therefore the expansion ratio is also increased as the engine load is lowered. At this time as well, the throttle valve 17 is kept fully open or substantially fully open, so the amount of intake air supplied into the combustion chamber 5 changes the closing timing of the intake valve 7 regardless of the throttle valve 17. Is controlled by that. At this time as well, the throttle valve 17 is kept fully open or substantially fully open, so that the pumping loss is substantially 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 reduced in proportion to the reduction 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 engine load is reduced to L ₁ , 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 held at the limit mechanical compression ratio in a region where the load is lower than the engine load L ₁ when the mechanical compression ratio reaches the limit mechanical compression ratio. Therefore, in the region where the load is lower than the engine load L ₁ , the mechanical compression ratio is maximum and the expansion ratio is also maximum. In other words, the mechanical compression ratio is maximized so that the maximum expansion ratio is obtained in a region where the load is lower than the engine load L ₁ .

When the engine load is further reduced from the engine load L ₁ when the mechanical compression ratio reaches the limit mechanical compression ratio, the closing timing of the intake valve 7 is delayed while the mechanical compression ratio is maintained at the limit mechanical compression ratio. . Therefore, as the engine load from the engine load L ₁ becomes lower when the mechanical compression ratio reaches the limit mechanical compression ratio, actual compression ratio is caused to decrease.

Further, in the example shown by the broken line in FIG. 9, the closing timing of the intake valve 7 is delayed as the engine load becomes lower from the engine high load operation state, and the engine when the mechanical compression ratio reaches the limit mechanical compression ratio. When the engine load is reduced to L ₂ lower than the load L _1, the closing timing of the intake valve 7 becomes the limit closing timing at which the intake air amount supplied into the combustion chamber 5 can be controlled. When the closing timing of the intake valve 7 reaches the limit closing timing, the intake valve 7 is closed 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 limiting closing timing. The timing is held at the limit closing timing.

Thus, in the region where the load is lower than the engine load L ₂ when the valve closing timing of the intake valve 7 reaches the limit valve closing timing, the mechanical compression ratio is maintained at the limit mechanical compression ratio, and the intake valve 7 The valve closing timing is held at the limit valve closing timing. Therefore, in this region, the actual compression ratio is kept almost constant. That is, in the example shown by the broken line in FIG. 9, the region where the load is higher than the engine load L ₁ when the mechanical compression ratio reaches the limit mechanical compression ratio, and the closing timing of the intake valve 7 are the limit closing timing. In the regions where the load is lower than the engine load L ₂ when the engine load is reached, the actual compression ratio is kept substantially constant in each region, and the engine load is in the region between the engine load L ₁ and the engine load L _2. The actual compression ratio is lowered as the value decreases.

On the other hand, if 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 the change in the closing timing of the intake valve 7. In the example shown by the broken line in FIG. 9, 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 17 causes the inside of the combustion chamber 5 to The amount of intake air supplied to is controlled. However, if the intake air amount is controlled by the throttle valve 17, the pumping loss increases.

If the intake air amount is controlled by the throttle valve 17, the pumping loss increases. Therefore, when the closing timing of the intake valve 7 reaches the limit closing timing so that such a pumping loss does not occur. In a region where the load is lower than the engine load L _{1, the} air-fuel ratio can be increased as the engine load decreases while the throttle valve 17 is kept fully open or substantially fully open. At this time, it is preferable to arrange the fuel injection valve 13 in the combustion chamber 5 and perform stratified combustion.

In the example described above, the engine load L ₁ when the mechanical compression ratio reaches the limit mechanical compression ratio is higher than the engine load L ₂ when the intake valve 7 closes to the limit closing timing. However, depending on the configuration of the variable compression ratio mechanism A and the variable valve timing mechanism B, the engine load L ₁ when the mechanical compression ratio reaches the limit mechanical compression ratio is different from the closing timing of the intake valve 7. In some cases, the engine load L ₂ may be lower. In this case, the closing timing of the intake valve 7 is between the engine load L ₂ when the closing timing of the intake valve 7 reaches the limit closing timing and the engine load L ₁ when the mechanical compression ratio reaches the limiting mechanical compression ratio. As the engine load decreases in a state where is kept at the limit valve closing timing, only the mechanical compression ratio is increased. For this reason, the engine load L ₂ when the closing timing of the intake valve 7 reaches the limit closing timing and the engine load L ₁ when the mechanical compression ratio reaches the limit mechanical compression ratio, the actual compression ratio is the engine load. It increases as it gets lower. Alternatively, in the 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 closing timing of the intake valve 7 is held at the limit closing timing and the mechanical compression ratio is also increased. You may make it hold | maintain substantially constant. In this case, the actual compression ratio is maintained substantially constant 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.

  Further, when the engine speed increases, the air-fuel mixture in the combustion chamber 5 is disturbed, so that knocking is less likely to occur. Therefore, in the embodiment according to the present invention, the actual compression ratio increases as the engine speed increases.

Further, 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 as long as 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.
In the example shown by the broken line 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.

  By the way, as described above, in the basic control of the spark ignition type internal combustion engine, the intake air amount supplied into the combustion chamber 5 is controlled by the closing timing of the intake valve 7. Accordingly, in the engine low load operation region, the closing timing of the intake valve 7 is retarded so as to approach the compression top dead center in order to reduce the amount of intake air supplied into the combustion chamber 5. Thus, when the valve closing timing of the intake valve 7 is retarded, the temperature of the intake gas supplied into the combustion chamber 5 increases. Hereinafter, the reason will be described with reference to FIG.

  FIG. 10 is a diagram schematically showing the flow of intake gas when the closing timing of the intake valve 7 is retarded. Even when the closing timing of the intake valve 7 is retarded, the intake valve 7 is opened during the intake stroke, that is, when the piston 4 is lowered. When the intake valve 7 is opened while the piston 4 is descending, the intake gas flows into the combustion chamber 5 from the intake port 8 as the piston 4 descends, as indicated by the white arrow in FIG. I'm damned.

  On the other hand, when the closing timing of the intake valve 7 is retarded, the intake valve 7 is opened for a certain period even during the compression stroke, that is, when the piston 4 is raised. If the intake valve 7 is opened while the piston 4 is raised, a part of the intake gas that has once flowed into the combustion chamber 5 is part of the combustion chamber 5 as shown by the white arrow in FIG. The air is blown back into the intake port 8 from the inside. Thus, the amount of the intake gas blown back from the combustion chamber 5 to the intake port 8 is such that the longer the period during which the intake valve 7 is opened while the piston 4 is raised, that is, the closing timing of the intake valve 7 is longer. The slower you go, the more.

  Further, as indicated by a black arrow in FIG. 10A, the intake gas that has flowed into the combustion chamber 5 receives heat from the cylinder wall surface that defines the combustion chamber 5. For this reason, the temperature of the intake gas flowing into the combustion chamber 5 is raised. Thereafter, a portion of the intake gas whose temperature has been raised in the combustion chamber 5 is blown back from the combustion chamber 5 into the intake port 8 during the compression stroke as described above. For this reason, the temperature of the intake gas once flowing into the combustion chamber 5 and blown back into the intake port 8 is higher than the temperature of the intake gas flowing into the combustion chamber 5. In particular, as the closing timing of the intake valve 7 is delayed, the amount of intake gas blown back from the combustion chamber 5 to the intake port 8 increases, so that the temperature of the intake gas in the intake port 8 increases. .

  Thus, when the temperature of the intake gas in the intake port 8 becomes high, the temperature of the intake gas supplied into the combustion chamber 5 in the next cycle is increased. Therefore, in summary, it can be said that the later the closing timing of the intake valve 7, the higher the temperature of the intake gas supplied into the combustion chamber 5.

  On the other hand, when the temperature of the intake gas supplied into the combustion chamber 5 increases, the temperature of the intake gas in the combustion chamber 5 when the piston 4 reaches the compression top dead center (hereinafter, “compression end temperature”). Will rise). Such a rise in the compression end temperature causes knocking. For this reason, when the compression end temperature rises, it is necessary to suppress the occurrence of knocking. As a technique for suppressing the occurrence of knocking, for example, it is conceivable to retard the ignition timing. However, retarding the ignition timing results in a decrease in thermal efficiency and a deterioration in fuel consumption.

  Therefore, in the embodiment of the present invention, the temperature of the intake gas supplied into the combustion chamber 5 is detected or estimated, and when the detected or estimated intake gas temperature is high, the valve closing timing of the intake valve 7 is advanced. At the same time, the mechanical compression ratio is reduced.

  FIG. 11 shows the temperature (intake air temperature) of the intake gas supplied into the combustion chamber 5 at a certain engine load, the closing timing of the intake valve 7, the mechanical compression ratio, the actual compression ratio, the throttle opening, and the intake air amount. It is a figure which shows the relationship. As shown in FIG. 11, the closing timing of the intake valve 7 is held at a constant timing on the retard side when the intake air temperature is equal to or lower than a predetermined reference temperature Tr. On the other hand, when the intake air temperature becomes higher than the reference temperature Tr, the closing timing of the intake valve 7 is advanced compared to when the intake air temperature is equal to or lower than the reference temperature Tr. In particular, in the present embodiment, in a region where the intake air temperature is higher than the reference temperature Tr, the closing timing of the intake valve 7 is advanced as the intake air temperature increases.

  Further, as shown in FIG. 11, the mechanical compression ratio is kept constant in a high state when the intake air temperature is equal to or lower than the reference temperature Tr. On the other hand, when the intake air temperature becomes higher than the reference temperature Tr, the mechanical compression ratio is made lower than when the intake air temperature is equal to or lower than the reference temperature Tr. In particular, in the present embodiment, in a region where the intake air temperature is higher than the reference temperature Tr, the mechanical compression ratio is lowered as the intake air temperature increases. In the present embodiment, the actual compression ratio is kept substantially constant even when the intake air temperature changes. Conversely, the mechanical compression ratio is lowered as the closing timing of the intake valve 7 is advanced so that the actual compression ratio is maintained substantially constant.

  Furthermore, as shown in FIG. 11, the opening degree of the throttle valve 17 is maintained in a large state when the intake air temperature is equal to or lower than the reference temperature Tr. On the other hand, when the intake air temperature becomes higher than the reference temperature Tr, the opening degree of the throttle valve 17 is made smaller than when the intake air temperature is lower than the reference temperature Tr. In particular, in the present embodiment, in the region where the intake air temperature is higher than the reference temperature Tr, the opening degree of the throttle valve 17 is made smaller as the intake air temperature becomes higher. In the present embodiment, the amount of intake air supplied to the combustion chamber 5 is kept substantially constant even if the intake air temperature changes. In other words, the opening of the throttle valve 17 is made smaller as the closing timing of the intake valve 7 is advanced so that the amount of intake air supplied to the combustion chamber 5 is kept substantially constant.

  Hereinafter, the operation and effect of controlling the valve closing timing of the intake valve 7, the mechanical compression ratio, and the opening of the throttle valve 17 according to the intake air temperature as described above will be described. As described above, when the intake air temperature becomes high, knocking is likely to occur, resulting in a decrease in thermal efficiency and a deterioration in fuel consumption. Here, in this embodiment, the valve closing timing of the intake valve 7 is advanced as the intake air temperature increases. When the closing timing of the intake valve 7 is advanced, the amount of intake gas once flowing into the combustion chamber 5 and then blown back into the intake port 8 is reduced, and the intake temperature is lowered accordingly. Therefore, according to the present embodiment, it is possible to suppress an increase in the intake air temperature by excessively retarding the closing timing of the intake valve 7.

  Further, if the mechanical compression ratio is kept constant when the closing timing of the intake valve 7 is advanced, the actual compression ratio is increased. However, when the actual compression ratio increases, the temperature (compression end temperature) and pressure (compression end pressure) in the combustion chamber 5 when the piston 4 reaches the compression top dead center will increase as a result. It will be generated. Here, according to the present embodiment, since the mechanical compression ratio is lowered when the closing timing of the intake valve 7 is advanced, the actual compression ratio is increased even when the closing timing of the intake valve 7 is advanced. Can be prevented from greatly increasing, so that the occurrence of knocking can be suppressed.

  Further, as described above, in the present embodiment, the amount of intake air supplied into the combustion chamber 5 is controlled by the closing timing of the intake valve 7, so that if the closing timing of the intake valve 7 is advanced, the combustion chamber The amount of intake air supplied into 5 increases. However, if the amount of intake air supplied into the combustion chamber 5 increases, torque exceeding the torque corresponding to the engine load is generated, or the air-fuel ratio changes greatly. Here, according to the present embodiment, since the opening degree of the throttle valve 17 is reduced when the closing timing of the intake valve 7 is advanced, even if the closing timing of the intake valve 7 is advanced. A large increase in the amount of intake air supplied into the combustion chamber 5 can be prevented, and hence torque fluctuations and deterioration of the air-fuel ratio can be suppressed.

  In this embodiment, only when the intake air temperature becomes higher than the reference temperature Tr, the closing timing of the intake valve 7, the mechanical compression ratio, and the opening degree of the throttle valve 17 are controlled according to the change in the intake air temperature. ing. This is because when the intake air temperature is equal to or lower than the reference temperature Tr, the decrease in thermal efficiency accompanying the increase in the intake air temperature is small.

  That is, as described above, when the intake air temperature rises, the ignition timing is retarded to avoid knocking associated with the rise in the compression end temperature. If the ignition timing is retarded in this way, the theoretical thermal efficiency is reduced. On the other hand, as can be seen from FIG. 7, the theoretical thermal efficiency is also lowered by reducing the mechanical compression ratio (ie, the expansion ratio). Here, when the intake air temperature is equal to or lower than the reference temperature Tr, the amount of decrease in the theoretical thermal efficiency due to the decrease in the expansion ratio is larger than the amount of decrease in the theoretical thermal efficiency due to the increase in the intake air temperature. For this reason, in such a temperature range, even if the intake air temperature rises, the theoretical thermal efficiency can be increased by keeping the expansion ratio high without lowering the expansion ratio, and thus fuel efficiency can be improved. On the other hand, when the intake air temperature is higher than the reference temperature Tr, the amount of decrease in the theoretical thermal efficiency due to the increase in the intake air temperature is larger than the amount of decrease in the theoretical heat efficiency due to the decrease in the expansion ratio. Therefore, in such a temperature range, when the intake air temperature rises, even if the expansion ratio is lowered, lowering the intake air temperature can increase the theoretical thermal efficiency, thereby improving the fuel efficiency.

  The solid line in FIG. 9 has constant parameters (air temperature, engine cooling water temperature, EGR rate when an EGR mechanism is provided, etc.) that affect the intake air temperature except when the intake valve 7 is closed. When the closing timing of the intake valve 7 is controlled according to the intake temperature as described above under the conditions, the closing timing of the intake valve 7, the mechanical compression ratio, the expansion ratio, the actual compression ratio, the throttle valve 17 Each change according to the engine load of the opening degree and the intake air amount is shown.

As shown by the solid line in FIG. 9, when the control is performed according to the intake air temperature, when the engine load is low, the valve closing timing of the intake valve 7 is performing the basic control (broken line in the figure). It can be advanced a lot compared to. The reason why the valve closing timing of the intake valve 7 is greatly advanced in this manner is that, in the engine low load operation region, the degree of increase in the intake air temperature is large when the basic control is performed. This is because when the control is performed, the advance amount of the closing timing of the intake valve 7 is also increased in order to suppress an increase in the intake air temperature. On the other hand, in the region where the engine load is L ₂ or more, the advance amount of the closing timing of the intake valve 7 when the control is performed according to the intake air temperature when the basic control is performed is high when the engine load is high. The smaller it becomes. This is because even when the basic control is being performed, the valve closing timing of the intake valve 7 is advanced as the engine load increases, and the intake air temperature decreases accordingly.

Further, as shown by the solid line in FIG. 9, when the control according to the intake air temperature is performed, the mechanical compression ratio (and the expansion ratio) is performed in the basic control in the region where the engine load is low (in the figure). Lower than the broken line). This is because when the control according to the intake air temperature is being performed, the closing timing of the intake valve 7 is advanced compared to when the basic control is being performed. This is because the ratio does not increase (particularly, in the present embodiment, the actual compression ratio does not change). On the other hand, in the region where the engine load is L ₂ or more, the amount of decrease in the mechanical compression ratio when performing control according to the intake air temperature when performing basic control becomes smaller as the engine load increases. This is because the advance amount of the valve closing timing of the intake valve 7 when the control according to the intake air temperature is being performed becomes smaller as the engine load becomes higher.

Furthermore, as shown by the solid line in FIG. 9, when the control according to the intake air temperature is performed, the opening degree of the throttle valve 17 is performing the basic control in the region where the engine load is low (the broken line in the figure). ) Is made smaller. This is because when the control according to the intake air temperature is being performed, the closing timing of the intake valve 7 is advanced relative to when the basic control is being performed. This is to prevent the amount from increasing. On the other hand, in the region where the engine load is L ₂ or more, the amount of decrease in the opening of the throttle valve 17 with respect to the basic control when the control is performed according to the intake air temperature is increased as the engine load increases. Get smaller. This is because the advance amount of the valve closing timing of the intake valve 7 when the control according to the intake air temperature is being performed becomes smaller as the engine load becomes higher.

  In the present embodiment, the intake air temperature is measured by a temperature sensor (not shown) provided in the intake port 8. By providing the temperature sensor in the intake port 8 in this way, even if a portion of the intake gas once flowing into the combustion chamber 5 is blown back from the combustion chamber 5 into the intake port 8, The temperature of the intake gas in the port 8, that is, the temperature of the intake gas supplied into the combustion chamber 5 can be accurately detected.

  Alternatively, the temperature of the intake gas supplied into the combustion chamber 5 is calculated using a pressure sensor (not shown) for detecting the pressure in the combustion chamber 5 and the intake air amount detector 18 instead of the temperature sensor. You may do it. That is, the mass flow rate of the intake gas supplied into the combustion chamber 5 is known by the intake air amount detector 18, and therefore the mass of the intake gas filled in the combustion chamber 5 when the intake valve 7 is closed is known. In addition, the pressure of the intake gas in the combustion chamber 5 when the intake valve 7 is closed can be known from the pressure sensor. Furthermore, the volume of the combustion chamber 5 when the intake valve 7 is closed is known from the crank angle when the intake valve 7 is closed. Thus, by using a pressure sensor or the like, the mass, pressure, and volume of the intake gas in the combustion chamber 5 when the intake valve 7 is closed (or a predetermined timing after the intake valve 7 is closed) can be obtained. Therefore, based on these parameters, the temperature of the intake gas supplied into the combustion chamber 5 can be calculated by a state equation or the like.

  FIG. 12 is a flowchart showing a control routine for controlling the closing timing of the intake valve 7, the mechanical compression ratio, and the opening degree of the throttle valve 17 according to the present invention. As shown in FIG. 12, first, in step S11, the engine load L and the engine speed Ne are detected based on the outputs of the load sensor 41 and the crank angle sensor. Next, in step S12, based on the engine load L and the engine speed Ne detected in step S11, for example, using a map or the like stored in advance in the ROM 32 of the ECU 30, the basic as shown by the broken line in FIG. The target valve closing timing IVCter, the target mechanical compression ratio εmter, and the target throttle opening degree θtter of the intake valve 7 when performing control are calculated.

Then, in step S13, the detected engine load L at step S11, it is determined whether it is set load L ₃ or less. This set load L ₃ is an engine load that causes a large amount of intake air to be sucked into the combustion chamber 5 due to a high engine load, and therefore a small amount of intake gas blown back from the combustion chamber 5. The engine load is such that the super high expansion ratio cycle is not executed at least in a region where the load is higher than the engine load. In step S13, if the engine load L is determined to be higher than the set load L ₃ is not necessary to consider the temperature increase of the intake gas by blowback of the intake gas, step S14~S18 is skipped.

On the other hand, in step S13, if the engine load L is determined to set the load L ₃ or less, the process proceeds to step S14. In step S14, the pressure P of the intake gas in the combustion chamber 5 after the intake valve 7 is closed is detected by a pressure sensor (not shown) that detects the pressure P of the intake gas in the combustion chamber 5. Next, in step S15, the temperature T of the intake gas supplied into the combustion chamber 5 is calculated based on the pressure P in the combustion chamber 5 detected in step S14.

  Next, in step S16, it is determined whether or not the temperature T of the intake gas calculated in step S15 is higher than the reference temperature Tr. If it is determined in step S16 that the temperature T of the intake gas is equal to or lower than the reference temperature Tr, the temperature of the intake gas is not so high as to deteriorate the thermal efficiency of the internal combustion engine due to the blow-back of the intake gas. S17 to S18 are skipped. On the other hand, if it is determined in step S16 that the temperature T of the intake gas is higher than the reference temperature Tr, the process proceeds to step S17. In step S17, based on the difference ΔT (= T−Tr) between the intake gas temperature T calculated in step S15 and the reference temperature Tr, the intake valve 7 is closed using a map as shown in FIG. A timing correction amount ΔIVC, a compression ratio correction amount Δεm, and a throttle opening correction amount Δθt are calculated.

  Next, in step S18, the timing advanced by the valve closing timing correction amount ΔIVC from the target valve closing timing IVCter of the intake valve 7 calculated in step S12 is set as a new target valve closing timing IVCter (IVCter = IVCter−). ΔIVC). Further, a value obtained by subtracting the compression ratio correction amount Δεm from the target mechanical compression ratio εmter calculated in Step S12 is set as a new target mechanical compression ratio εmter (εmter = εmter−Δεm). Further, a value obtained by subtracting the throttle opening correction amount Δθt from the target throttle opening θtter calculated in step S12 is set as a new target throttle opening θtter (θtter = θtter−Δθt). Next, in step S19, the variable valve timing mechanism B is controlled so that the closing timing of the intake valve 7 becomes the target closing timing IVCter calculated in step S12 or S18. Further, the variable compression ratio mechanism A is controlled so that the mechanical compression ratio becomes the target mechanical compression ratio εmter calculated in step S12 or 18. Further, the throttle valve driving actuator 16 is controlled so that the opening degree of the throttle valve 17 becomes the target throttle opening degree θt calculated in step S12 or S18.

  Since the temperature of the cylinder block 2 and the like is low when the internal combustion engine is cold started, the temperature of the intake gas does not rise so much even if the intake gas once flowing into the combustion chamber 5 is blown back into the intake port 8. . For this reason, when the internal combustion engine is cold started, the intake gas temperature is regarded as lower than the reference temperature Tr without detecting or estimating the intake gas temperature, and the valve closing timing of the intake valve 7 is controlled. It may be.

Next, a second embodiment of the present invention will be described with reference to FIGS. In the above-described embodiment shown in FIG. 9, the actual compression ratio when the control according to the intake air temperature is performed is kept the same as the actual compression ratio when the basic control is performed in all load regions. . However, in this embodiment, in the region where the load is lower than the engine load L ₁ when the mechanical compression ratio reaches the limit mechanical compression ratio, the actual compression ratio when the control according to the intake air temperature is performed is the basic It is assumed that the actual compression ratio is higher when the control is performed.

FIG. 14 shows the temperature (intake air temperature) of the intake gas supplied into the combustion chamber 5 at a certain engine load having a load lower than the engine load L ₁ when the control according to the intake air temperature is performed in the present embodiment. 12 is a view similar to FIG. 11, showing the relationship between the intake valve 7 and the closing timing of the intake valve 7. As can be seen from FIG. 14, in the present embodiment as well, as in the case of FIG. 11, when the intake air temperature becomes higher than the reference temperature Tr, the intake valve 7 is closed as the intake air temperature increases. The valve timing is advanced. Similarly, as in the case of FIG. 11, when the intake air temperature becomes higher than the reference temperature Tr, the mechanical compression ratio is lowered as the intake air temperature increases. However, the extent to which the mechanical compression ratio decreases as the intake air temperature increases is smaller than in the case of FIG.

On the other hand, unlike the case of FIG. 11, the actual compression ratio is maintained constant when the intake air temperature is equal to or lower than the reference temperature Tr. However, when the intake air temperature is higher than the reference temperature Tr, the actual compression ratio is Compared to high. In particular, in this embodiment, in a region where the intake air temperature is higher than the reference temperature Tr, the actual compression ratio is increased as the intake air temperature increases. In the region of a load higher than the engine load L _1, the closing timing of the intake valve 7, the mechanical compression ratio and actual compression ratio is controlled as in the example shown in FIG. 11 according to the intake air temperature.

  FIG. 15 shows each change of the closing timing of the intake valve 7 according to the engine load when the closing timing of the intake valve 7 according to the intake air temperature in the present embodiment is controlled. It is a simple figure. In particular, the broken lines in FIG. 15 indicate changes in each parameter when the basic control is performed, while the solid lines in FIG. 15 indicate parameters that affect the intake air temperature other than the closing timing of the intake valve 7. The change of each parameter at the time of controlling valve closing timing etc. of the intake valve 7 according to the intake air temperature in the present embodiment under the condition that is constant is shown.

That is, as described above, when the basic control is performed, in the region where the load is higher than the engine load L ₁ when the mechanical compression ratio reaches the limit mechanical compression ratio, the actual compression ratio is within a range where knocking does not occur. It is kept at the limit compression ratio. However, in the region where the load is lower than the engine load L ₁ when the mechanical compression ratio reaches the limit mechanical compression ratio, the closing timing of the intake valve 7 is not increased even though the mechanical compression ratio cannot be increased further. Since it is delayed, the actual compression ratio is made lower than in a region where the load is higher than the engine load L ₁ . For this reason, in the region where the load is lower than the engine load L ₁ , a margin can be provided for the limit of knocking.

On the other hand, as described above, when the control according to the intake air temperature is performed, the closing timing of the intake valve 7 is advanced and the mechanical compression ratio is lowered as compared with the case where the basic control is performed. For this reason, even if the engine load becomes L ₁ , the mechanical compression ratio does not reach the limit mechanical compression ratio. Therefore, the mechanical compression ratio can be further increased even in a region where the load is lower than L ₁ . Therefore, in this embodiment, so as to increase the mechanical compression ratio as the engine load is also the engine load becomes lower in the region lower than L _1. Thereby, in the region where the engine load is lower than L _1, the actual compression ratio when the control according to the intake air temperature is performed is higher than the actual compression ratio when the basic control is performed. The

Incidentally, in the illustrated embodiment, if the control is performed in accordance with the intake air temperature, actual compression ratio engine load is in the region lower than L ₁ are the same as the actual compression ratio engine load is in the region higher than L ₁ In addition, it is controlled to be constant. However, the basic control is performed so that the actual compression ratio in the region lower than the engine load L ₁ when performing control according to the intake air temperature is lower than the actual compression ratio in the region where the engine load is higher than L _1. It may be controlled so as to be higher than the actual compression ratio when performing the above.

1 Crankcase 2 Cylinder block 3 Cylinder head 4 Piston 5 Combustion chamber 7 Intake valve 8 Intake port A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (6)

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 the temperature of the intake gas supplied into the combustion chamber when the intake valve is closed is detected or estimated An intake gas temperature detecting means for
During engine low load operation, when the temperature detected or estimated by the intake gas temperature detection means is higher than the reference temperature, the closing timing of the intake valve is closer to the compression bottom dead center than when the temperature is lower than the reference temperature. A spark ignition internal combustion engine that is advanced and has a reduced mechanical compression ratio.   The throttle valve opening is made smaller when the temperature detected or estimated by the intake gas temperature detection means is higher than the reference temperature during engine low load operation compared to when the temperature is lower than the reference temperature. 2. The spark ignition internal combustion engine according to 1.   The spark ignition internal combustion engine according to claim 1 or 2, wherein the actual compression ratio is made substantially the same regardless of the temperature detected or estimated by the intake gas temperature detection means during engine low load operation.   During engine low load operation, when the temperature detected or estimated by the intake gas temperature detection means is higher than the reference temperature, the closing timing of the intake valve is set to the compression bottom dead center as the detected or estimated temperature increases. The spark ignition internal combustion engine according to any one of claims 1 to 3, wherein the spark ignition type internal combustion engine is advanced so as to approach, the mechanical compression ratio is reduced, and the opening of the throttle valve is reduced.   During engine low load operation, when the temperature detected or estimated by the intake gas temperature detection means is equal to or lower than the reference temperature, the intake valve closing timing and the mechanical compression ratio are made substantially constant regardless of the detected or estimated temperature. The spark ignition internal combustion engine according to any one of claims 1 to 4, wherein the spark ignition internal combustion engine is maintained.   6. When the internal combustion engine is cold-started, the intake valve closing timing and the mechanical compression ratio are controlled on the assumption that the temperature detected or estimated by the intake gas temperature detecting means is lower than the reference temperature. The spark ignition internal combustion engine according to any one of the above.

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