JP5589707B2 - Control device for internal combustion engine - Google Patents

Description

  The present invention relates to a control device for an internal combustion engine.

  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, and at the time of engine low load operation, the mechanical compression ratio is increased and the closing timing of the intake valve The engine is under low load operation by retarding the engine from the intake bottom dead center, lowering the mechanical compression ratio during high engine load operation, and advancing the intake valve close timing closer to the intake bottom dead center There is known a spark ignition type internal combustion engine in which the actual compression ratio is substantially the same between the engine and the high engine load operation (for example, Patent Document 1).

JP 2007-303423 A JP 2005-127229 A JP 2005-133604 A

  By the way, in the internal combustion engine described in Patent Document 1, the mechanical compression ratio and the closing timing of the intake valve are controlled so that the actual compression ratio is substantially the same between the engine low load operation and the engine high load operation. . This is because while increasing the actual compression ratio increases the thermal efficiency, knocking occurs when the actual compression ratio increases, so the actual compression ratio should be kept as high as possible within the range where knocking does not occur. It is a thing.

  Here, the actual compression ratio is calculated based on the combustion chamber volume at the compression operation start timing when the compression operation is actually started. In the internal combustion engine described in Patent Document 1, the actual compression ratio is calculated using the closing timing of the intake valve as the compression action start timing. However, in an actual internal combustion engine, the compression action of the air-fuel mixture by the piston is started before the intake valve is closed. Therefore, if the intake valve closing timing is approximated to the compression action start timing, the actual compression ratio is accurately determined. I can't ask for it.

  Further, the compression action start time is used not only to calculate the above-described actual compression ratio but also to calculate various parameters of the internal combustion engine. Examples of such parameters include the temperature (compression end temperature) and pressure (compression end pressure) in the combustion chamber when the piston is at compression top dead center, the temperature and pressure in the combustion chamber during the compression stroke, and the like. It is done. Even in this case, if the valve closing timing of the intake valve is approximated with the compression operation start timing, these parameters cannot be obtained accurately.

  Accordingly, in view of the above problems, an object of the present invention is to accurately obtain the compression action start time and to obtain the obtained compression action start time in an internal combustion engine having a variable valve timing mechanism that can change the closing timing of the intake valve. The present invention provides a control device that appropriately controls an internal combustion engine based on the above.

In order to solve the above-mentioned problem, in the first invention, in the control device for an internal combustion engine comprising a variable valve timing mechanism capable of changing the closing timing of the intake valve, which is the timing when the lift amount of the intake valve becomes zero, Based on engine operating parameters including at least one of the engine speed, intake valve closing timing, and intake valve lift during the intake valve opening period, the combustion chamber volume and piston at the closing timing of the intake valve A deviation amount calculating means for calculating a deviation amount from the combustion chamber volume at the compression action start timing at which the actual compression action starts, and a valve closing time volume calculating means for calculating the combustion chamber volume at the actual valve closing timing of the intake valve And when the compression action is started by subtracting the deviation amount calculated by the deviation amount calculating means from the combustion chamber volume at the actual intake valve closing timing calculated by the valve closing volume calculating means. Comprising: a compression start volume calculating means for calculating a combustion chamber volume, and a control parameter setting means for setting a control parameter of the internal combustion engine based on the combustion chamber volume at the compression action start timing calculated by the compression start volume calculating means in The divergence amount calculation means calculates the divergence amount by correcting a predetermined reference divergence amount based on the closing timing of the intake valve .

In the second invention, in the first invention, the deviation amount calculating means, the reference deviation amount to a predetermined further comprises at least one of the lift amount of the intake valve in the engine speed及 Beauty intake valve during the valve open period The deviation amount is calculated by correcting based on one of them.

  According to a third aspect, in the second aspect, the reference divergence amount assumes that the compression of the intake gas in the combustion chamber after the intake valve is closed is adiabatic compression after the intake valve is closed in a predetermined engine operating state. When it is assumed that the compression is also performed before the intake valve is closed, the combustion chamber volume at a time when the pressure in the combustion chamber becomes a predetermined pressure lower than the pressure in the combustion chamber at the intake valve closing timing; This is the difference from the combustion chamber volume at the closing timing of the intake valve.

  According to a fourth aspect, in the third aspect, the predetermined pressure is the intake pipe pressure in the predetermined engine operating state.

  In the fifth invention, in any one of the second to fourth inventions, the correction amount with respect to the reference deviation amount when calculating the deviation amount is increased as the engine speed increases.

  In the sixth invention, in any one of the second to fifth inventions, the correction amount with respect to the reference deviation amount when calculating the deviation amount is such that the lift amount of the intake valve during the intake valve opening period decreases. Increased.

According to a seventh invention, in any one of the first to sixth inventions, a variable compression ratio mechanism capable of changing a mechanical compression ratio is further provided, and the control parameter of the internal combustion engine is a mechanical compression ratio.

In an eighth invention, in any one of the first to sixth inventions, the control parameter of the internal combustion engine is at least one of a fuel injection timing from a fuel injection valve or an ignition timing by a spark plug.

  According to the present invention, the compression action start timing can be accurately obtained, and the internal combustion engine can be appropriately controlled based on the obtained compression action start timing.

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 of an intake 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. FIG. 3 is a double logarithmic PV diagram showing transition of the combustion chamber volume and the combustion chamber pressure during the compression stroke. It is a figure which shows the relationship between intake valve closing timing, engine speed, and deviation | shift amount. It is a figure which shows the relationship between intake valve closing timing, lift amount, and deviation | shift amount. It is a figure which shows the map of the correction coefficient according to an engine speed etc. It is a flowchart which shows the setting procedures, such as a target mechanical compression ratio.

  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 air flow meter 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 21 containing an exhaust purification catalyst 20 via an exhaust manifold 19, and an air-fuel ratio sensor 22 is disposed in the exhaust manifold 19. As the exhaust purification catalyst 20, any catalyst such as a three-way catalyst, a NO _x storage reduction catalyst, and a NO _x selective reduction catalyst may be used as long as unburned HC, CO, and NO _x in the exhaust gas can be purified.

  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 axis direction at the connecting portion between the crankcase 1 and the cylinder block 2. There is provided a variable compression ratio mechanism A capable of changing the volume of the combustion chamber 5 at the time, and an actual compression action start timing changing mechanism B capable of changing the actual start time of the compression action. In the embodiment shown in FIG. 1, the actual compression operation start timing changing mechanism B is composed of a variable valve timing mechanism capable of controlling the closing timing of the intake valve 7.

  The electronic control unit 30 is composed of a digital computer, and is connected to each other by a bidirectional bus 31. A ROM (read only memory) 32, a RAM (random access memory) 33, a CPU (microprocessor) 34, an input port 35 and an output port 36. It comprises. Output signals of the air flow meter 18 and the air-fuel ratio sensor 22 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 depression amount L of the accelerator pedal 40 is connected to the accelerator pedal 40, and the output voltage of the load sensor 41 is input to the input port 35 via the corresponding AD converter 37. Is done. Further, a crank angle sensor 42 that generates an output pulse every time the crankshaft rotates, for example, 15 ° is connected to the input port 35. On the other hand, the output port 36 is connected to the spark plug 6, the fuel injection valve 13, the throttle valve driving actuator 16, the variable compression ratio mechanism A, and the variable valve timing mechanism B through corresponding drive circuits 38.

  2 is an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 1, and FIG. 3 is a side sectional view of the internal combustion engine schematically shown. Referring to FIG. 2, a plurality of 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 58 is rotatably inserted into each cam insertion hole 53. It is fixed. These circular cams 58 are coaxial with the rotational axes of the camshafts 54 and 55. On the other hand, as shown in FIG. 3, eccentric shafts 57 arranged eccentrically with respect to the rotational axes of the cam shafts 54 and 55 extend on both sides of each circular cam 58, and another circular cam is disposed on the eccentric shaft 57. 56 is mounted eccentrically and rotatable. As shown in FIG. 2, these circular cams 56 are arranged on both sides of each circular cam 58, and these circular cams 56 are rotatably inserted into the corresponding cam insertion holes 51.

  When the circular cams 58 fixed on the camshafts 54 and 55 are rotated in opposite directions as indicated by arrows in FIG. 3A from the state shown in FIG. The circular cam 56 rotates in the direction opposite to the circular cam 58 in the cam insertion hole 51 in order to move in the away direction, and as shown in FIG. Position. Next, when the circular cam 58 is further rotated in the direction indicated by the arrow, the eccentric shaft 57 is at the lowest position as shown in FIG.

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

  3A to 3C, the relative positions of the crankcase 1 and the cylinder block 2 are determined by the distance between the center a of the circular cam 58 and the center c of the circular cam 56. The cylinder block 2 moves away from the crankcase 1 as the distance between the center a of the cylinder and the center c of the circular cam 56 increases. That is, the variable compression ratio mechanism A changes the relative position between the crankcase 1 and the cylinder block 2 by a crank mechanism using a rotating cam. When the cylinder block 2 moves away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is 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 worms 61 and 62 having opposite spiral directions are attached to the rotation shaft of the drive motor 59, respectively. Worm wheels 63 and 64 meshing with the worms 61 and 62 are fixed to 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, the cam phase changer B <b> 1, and the camshaft 70 and the intake valve 7 valve lifter. 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 according to the present invention, any type of mechanism may be used as long as it is a variable valve closing timing mechanism capable of changing 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. 6A, 6B, and 6C show an engine having a combustion chamber volume of 50 ml and a piston stroke volume of 500 ml for the sake of explanation. , (B), (C), the combustion chamber volume represents the volume of the combustion chamber when the piston is located at the compression top dead center.

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

  FIG. 6B describes the actual compression ratio. This actual compression ratio is a value determined from the actual piston stroke volume and the combustion chamber volume from when the compression action is actually started until the piston reaches top dead center, and this actual compression ratio is (combustion chamber volume + actual (Stroke volume) / combustion chamber volume. That is, as shown in FIG. 6B, even if the piston starts to rise in the compression stroke, the compression action is not performed while the intake valve is open, and actual compression is performed as 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 super expansion ratio cycle used in the present invention will be described with reference to FIGS. FIG. 7 shows the relationship between the theoretical thermal efficiency and the expansion ratio, and FIG. 8 shows a comparison between a normal cycle and an ultrahigh expansion ratio cycle that are selectively used according to the load in the present invention.

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

  The solid line in FIG. 7 shows the change in the theoretical thermal efficiency when the actual compression ratio and the expansion ratio are substantially equal, that is, in a normal cycle. In this case, it can be seen that the theoretical thermal efficiency increases as the expansion ratio increases, that is, as the actual compression ratio increases. Therefore, in order to increase the theoretical thermal efficiency in a normal cycle, 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, under such circumstances, it is considered to increase the theoretical thermal efficiency while strictly dividing the mechanical compression ratio and the actual compression ratio. As a result, the theoretical thermal efficiency is governed by the expansion ratio, and the actual compression ratio is compared to the theoretical thermal efficiency. Was found to have little effect. That is, if the actual compression ratio is increased, the explosive force is increased, but a large amount of energy is required for compression. Thus, even if the actual compression ratio is increased, the theoretical thermal efficiency is hardly increased.

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

  Thus, if the actual compression ratio is maintained at a low value, knocking does not occur. Therefore, if the expansion ratio is increased while the actual compression ratio is maintained at a low value, the theoretical thermal efficiency is reduced while preventing the occurrence of knocking. Can 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 to this case, only the expansion ratio is shown in FIG. 8B. It can be seen that it has been increased to 26. Therefore, such a cycle is referred to as an ultra-high expansion ratio cycle.

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

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

Next, the overall operation control will be schematically described with reference to FIG.
FIG. 9 shows changes in the intake air amount, the intake valve closing timing, the mechanical compression ratio, the expansion ratio, the actual compression ratio, and the opening degree of the throttle valve 17 according to the engine load at a certain engine speed. . 9 shows that the average air-fuel ratio in the combustion chamber 5 is an output signal of the air-fuel ratio sensor 22 so that unburned HC, CO and NO _x in the exhaust gas can be simultaneously reduced by the exhaust purification catalyst in the catalytic converter 21. This shows a case where feedback control is performed to the theoretical air-fuel ratio based on the above.

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

  On the other hand, as shown by the solid 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, as shown in FIG. 9, the mechanical compression ratio is increased as the engine load is lowered so that the actual compression ratio is kept substantially constant. Therefore, the expansion ratio is also increased as the engine load is lowered. At this time as well, the throttle valve 17 is held in a fully open state or almost fully open state, and therefore the amount of intake air supplied into the combustion chamber 5 is mainly closed without depending on the throttle valve 17. It is controlled by changing the time.

  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, in the example shown in FIG. 9, the air-fuel ratio in the combustion chamber 5 is the stoichiometric air-fuel ratio, so the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is proportional to the fuel amount. Will change.

When the engine load is further reduced, the mechanical compression ratio is further increased, and when the engine load is lowered to the medium load L ₁ slightly close to the low load, the mechanical compression ratio reaches a limit mechanical compression ratio that is a structural limit of the combustion chamber 5. When the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is 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, the mechanical compression ratio is maximized and the expansion ratio is maximized at the time of low engine load operation and low engine load operation, that is, on the engine low load operation side. In other words, the mechanical compression ratio is maximized so that the maximum expansion ratio is obtained on the engine low load operation side.

On the other hand, in the embodiment shown in FIG. 9, when the engine load decreases to L ₁ , the closing timing of the intake valve 7 becomes the limit closing timing that can control the amount of intake air supplied into the combustion chamber 5. When the closing timing of the intake valve 7 reaches the limit closing timing, the closing timing of the intake valve 7 in a region where the load is lower than the engine load L ₁ when the closing timing of the intake valve 7 reaches the closing timing. Is held at the limit valve closing timing.

When the closing timing of the intake valve 7 is held at the limit closing timing, the amount of intake air can no longer be controlled by changing the closing timing of the intake valve 7. In the embodiment shown in FIG. 9, at this time, that is, in a region where the load is lower than the engine load L ₁ when the closing timing of the intake valve 7 reaches the limit closing timing, the throttle valve 17 supplies the fuel into the combustion chamber 5. The amount of intake air to be controlled is controlled, and the opening degree of the throttle valve 17 is made smaller as the engine load becomes lower.

On the other hand, as shown by the broken line in FIG. 9, the intake air amount can be controlled without depending on the throttle valve 17 by advancing the closing timing of the intake valve 7 as the engine load becomes lower. Accordingly, when expressing the case shown in FIG. 9 so as to include both the case indicated by the solid line and the case indicated by the broken line, in the embodiment according to the present invention, the valve closing timing of the intake valve 7 becomes smaller as the engine load becomes lower. It is moved in a direction away from the intake bottom dead center until a limit valve closing timing L ₁ that can control the amount of intake air supplied into the combustion chamber. As described above, the intake air amount can be controlled by changing the closing timing of the intake valve 7 as shown by the solid line in FIG. 9 or by changing it as shown by the broken line. Hereinafter, the present invention will be described by taking as an example a case where the closing timing of the intake valve 7 is changed as shown by a solid line in FIG.

  In the above embodiment, the valve closing timing and the mechanical compression ratio of the intake valve 7 are controlled so that the actual compression ratio becomes substantially constant regardless of the engine load, but the intake valve 7 is controlled based on other parameters. The valve closing timing and the mechanical compression ratio may be controlled. For example, from the viewpoint of suppressing the occurrence of knocking, the temperature of the air-fuel mixture in the combustion chamber 5 (compression) when the piston 4 is at the compression top dead center at least in a region where the engine load is greater than a certain engine load. The valve closing timing and the mechanical compression ratio of the intake valve 7 may be controlled so that the end temperature) or pressure (compression end pressure) becomes substantially constant.

  By the way, as explained with reference to FIG. 6B, the actual compression ratio is the combustion chamber volume when the piston is located at the compression top dead center from the combustion chamber volume when the compression action is actually started. This is the subtracted volume (hereinafter, “combustion chamber volume” means the volume in the combustion chamber 5 that changes as the piston moves). Therefore, in order to accurately calculate the actual compression ratio, it is necessary to accurately determine the combustion chamber volume when the compression action is actually started.

  Here, when the intake valve 7 is closed during the compression stroke, the intake gas once filled in the combustion chamber 5 as the piston 4 rises during the compression stroke until the intake valve 7 is closed. Since it is returned to the intake port 8, it is considered that the intake gas in the combustion chamber 5 is not compressed by the piston 4 until the intake valve 7 is closed. However, in practice, as can be seen from FIG. 5, the lift of the intake valve 7 is not constant while the intake valve 7 is open, and gradually decreases as the intake valve 7 approaches the valve closing timing. Become. Therefore, immediately before the intake valve 7 is closed, the intake valve 7 is throttled with respect to the intake gas returned from the combustion chamber 5 into the intake port 8, and as a result, the compression of the intake gas in the combustion chamber 5 is a fact. In addition, the operation is started before the intake valve 7 is closed.

  In general, an intake system of an internal combustion engine is designed such that inertia supercharging is performed by intake pulsation generated in the intake port 8 when the intake valve 7 is opened in a predetermined engine operating state. When inertia supercharging is performed in this manner, the pressure of the intake gas in the intake port 8 is high while the intake valve 7 is open during the compression stroke. For this reason, even if the piston 4 moves up, the intake gas in the combustion chamber 5 does not sufficiently return to the intake port 8, and as a result, the compression of the intake gas in the combustion chamber 5 starts before the intake valve 7 is closed. Will be.

  FIG. 10 is a log-log PV diagram showing changes in the combustion chamber volume V and the combustion chamber pressure P during the compression stroke. Point a in FIG. 10 indicates the start of the compression stroke, point b indicates the end of the compression stroke, and point c indicates when the intake valve 7 is closed (that is, the lift amount of the intake valve 7 becomes zero). Is shown). As can be seen from FIG. 10, the combustion chamber pressure P starts to rise from the point d before the intake valve 7 is closed. Accordingly, the actual compression ratio is calculated by calculating the combustion chamber volume when the intake valve 7 is closed (point c) as the combustion chamber volume when the compression action is started (hereinafter referred to as “compression start volume”). The actual compression ratio will be lower than the actual actual compression ratio.

  On the other hand, the combustion chamber pressure P starts to rise from the point d, but the intake gas in the combustion chamber 5 is not adiabatically compressed from the point d to the point c. The extent to which P rises is small. For this reason, when the actual compression ratio is calculated using the combustion chamber volume when the combustion chamber pressure P starts to rise (point d) as the compression start volume, the calculated actual compression ratio is higher than the actual actual compression ratio. turn into.

  Further, as described above, when the valve closing timing and the mechanical compression ratio of the intake valve 7 are controlled so that the compression end temperature or the compression end pressure is substantially constant, the compression end temperature and the compression end pressure are determined based on the actual compression ratio. It is conceivable to calculate. However, as described above, actual compression is performed using the combustion chamber volume when the intake valve 7 is closed (point c) and the combustion chamber volume when the pressure P in the combustion chamber 5 starts increasing (point d) as the compression start volume. When the ratio is calculated, since the calculated actual compression ratio is not an accurate value, the compression end temperature and the compression end pressure cannot be calculated accurately.

  By the way, the change of the intake gas in the combustion chamber 5 after the intake valve 7 is closed can be approximated to adiabatic compression. When adiabatic compression is performed, the combustion chamber volume V and the combustion chamber pressure P change linearly in the log-log PV diagram. This is apparent from the fact that the point a and the point b in FIG. 10 are almost linear.

  A straight line obtained by extending the adiabatic compression line between the points ab (broken line M in the figure) is combustion when it is assumed that the intake gas in the combustion chamber 5 is adiabatically compressed before the intake valve 7 is closed. The relationship between the chamber volume V and the combustion chamber pressure P is represented. Also, the straight line M and the intake pipe pressure (the average pressure in the intake pipe downstream of the throttle valve 17 and upstream of the intake valve 7; or the average pressure in the surge tank 12; broken line N in the figure). The intersection (point e) represents the time when the compression of the intake gas in the combustion chamber 5 is started when it is assumed that the intake gas in the combustion chamber 5 is adiabatically compressed even before the intake valve 7 is closed. ing.

  Therefore, in this embodiment, the intersection (point e) between the straight line M and the straight line N is approximated to the time when the compression action is started, and the combustion chamber volume at this point e is combusted when the compression action is started. It is supposed to approximate the chamber volume (compression start volume). By approximating in this way, the compression start volume can be determined relatively accurately, and the actual compression ratio, compression end temperature, and compression end pressure can be determined relatively accurately. become.

Specifically, for each engine operating state, a deviation amount ΔV between the combustion chamber volume V ₁ when the intake valve 7 is closed and the combustion chamber volume V ₂ at the point e is calculated in advance by experiment or calculation and used as a map. It is stored in the ROM 32 of the ECU 30. During engine operation, the combustion chamber volume when the intake valve is closed is calculated based on the current valve closing timing of the intake valve 7, and the deviation amount ΔV is calculated based on the engine operation state. The compression start volume is calculated by subtracting the deviation amount ΔV from the combustion chamber volume when the intake valve is closed.

In other words, in this embodiment, it is assumed that the compression of the intake gas in the combustion chamber 5 after the intake valve 7 is closed is adiabatic compression, and the adiabatic compression is also performed before the intake valve is closed. Assuming that, the difference between the combustion chamber volume V ₂ when the pressure in the combustion chamber 5 becomes the intake pipe pressure and the combustion chamber volume V ₁ when the intake valve 7 is closed is calculated as the deviation amount ΔV. At the same time, the combustion chamber volume at the closing timing of the intake valve 7 is calculated, and the combustion chamber volume at the compression action start timing is calculated by subtracting the deviation amount ΔV from the calculated combustion chamber volume at the closing timing of the intake valve 7. Is to be calculated.

In the above embodiment, the difference between the combustion chamber volume V ₂ when the pressure in the combustion chamber 5 becomes the intake pipe pressure and the combustion chamber volume V ₁ when the intake valve 7 is closed is calculated as the deviation amount ΔV. doing. However, V ₂ does not necessarily have to be the volume of the combustion chamber at the time when the pressure in the combustion chamber 5 becomes the pressure in the intake pipe, and may be slightly different from the pressure in the intake pipe or may be atmospheric pressure. Good.

  Next, the relationship between the engine operating state and the deviation amount ΔV will be described. FIG. 11 is a diagram showing the relationship between the deviation amount ΔV, the closing timing of the intake valve 7 and the engine speed. As can be seen from FIG. 11, the deviation amount ΔV increases as the engine speed increases. This is because as the engine speed increases, the flow velocity of the intake gas when returning to the intake port 8 increases, and the influence of the throttle by the intake valve 7 increases.

  As can be seen from FIG. 11, the deviation amount ΔV also increases when the intake valve 7 closes at an intermediate timing (near 90 ° ABDC). That is, when the closing timing of the intake valve 7 is an intermediate timing, the rate of decrease in the combustion chamber volume per unit crank angle (and hence per unit time) increases. Therefore, a large amount of the intake gas once sucked into the combustion chamber 5 tends to return to the intake port 8 in the vicinity of the closing timing of the intake valve 7. However, in the vicinity of the closing timing of the intake valve 7, the lift of the intake valve 7 is small, so that the intake gas in the combustion chamber 5 cannot sufficiently return to the intake port 8. For this reason, even when the closing timing of the intake valve 7 is an intermediate timing, the deviation amount ΔV increases.

  Further, as can be seen from FIG. 11, the deviation amount ΔV is relatively large even when the closing timing of the intake valve 7 is the retarded timing. That is, when the closing timing of the intake valve 7 is a retarded timing, the time from the start of the compression stroke to the closing of the intake valve 7 becomes long. During this time, the intake gas once sucked into the combustion chamber 5 is returned to the intake port 8 over a long period of time. During this period, the intake valve 7 is throttled against the intake gas returned to the intake port 8. Therefore, if this period is long, the deviation amount ΔV increases. Therefore, even when the closing timing of the intake valve 7 is a retarded timing, the deviation amount ΔV is relatively large.

  As can be seen from FIG. 12, the deviation amount ΔV increases as the lift amount of the intake valve 7 during the opening period of the intake valve 7 decreases. That is, when the lift amount is small as indicated by c in FIG. 5B, the throttle when the intake gas once sucked into the combustion chamber 5 returns to the intake port 8 again becomes large. This is because when the lift amount is large as indicated by a in B), the throttle when the intake gas once sucked into the combustion chamber 5 returns to the intake port 8 again becomes small.

  As described above, the deviation amount ΔV changes according to the engine speed, the closing timing of the intake valve 7 and the lift amount of the intake valve 7 during the opening period of the intake valve 7. The above-described relationship between the parameter and the deviation amount ΔV is obtained in advance by experiment or calculation, and is stored in the ROM 32 of the ECU 30 as a map, and during engine operation, the deviation amount is obtained using the map based on the values of these operation parameters. ΔV is calculated.

Alternatively, the deviation amount ΔV may be calculated during engine operation as follows.
First, in a specific engine operating state, that is, the engine speed is a certain engine speed, the closing timing of the intake valve 7 is a certain closing timing, and the intake valve 7 is in the open valve opening period. In a state where the lift amount is a certain lift amount, a deviation amount ΔV between the combustion chamber volume V ₁ when the intake valve 7 is closed and the combustion chamber volume 5 V ₂ at the point e is calculated in advance by experiment or calculation. The deviation amount ΔV is stored in the ROM 32 of the ECU 30 as the reference deviation amount ΔV _base .

In addition, based on the relationship between the engine speed and the deviation amount ΔV, a correction coefficient (rotation speed correction coefficient) k _re corresponding to the engine speed is calculated in advance by experiment or calculation. As shown in FIG. 13A, the rotational speed correction coefficient k _re is small when the engine rotational speed is low, and is set to a value that increases as the engine rotational speed increases.

Further, based on the relationship between the valve closing timing of the intake valve 7 and the deviation amount ΔV, a correction coefficient (valve closing timing correction coefficient) k _ivc corresponding to the valve closing timing of the intake valve 7 is calculated in advance by experiment or calculation. . As shown in FIG. _{13B, the} valve closing timing correction coefficient k _ivc is set to a small value when the closing timing of the intake valve 7 is the advanced timing, and the closing timing of the intake valve 7 is medium. When the time is about the time or the time on the retard side, a large value is set.

Further, based on the relationship between the lift amount of the intake valve 7 during the valve opening period of the intake valve 7 and the deviation amount ΔV, a correction coefficient (lift) corresponding to the lift amount of the intake valve 7 during the valve opening period of the intake valve 7 is established. (Quantity correction coefficient) k _la is calculated in advance by experiment or calculation. As shown in FIG. 13C, the lift amount correction coefficient k _la is a value that increases when the lift amount of the intake valve 7 is small and decreases as the lift amount of the intake valve 7 increases. The

During engine operation, the correction coefficients k _re , k _ivc , and the like are based on the values of operation parameters such as the engine speed, the closing timing of the intake valve 7 and the lift amount of the intake valve 7 during the opening period of the intake valve 7. k _la is calculated, and the deviation amount ΔV is calculated by multiplying the reference deviation amount ΔV described above by the correction coefficients k _re , k _ivc , k _la (ΔV = ΔV _base · k _re · k _ivc · k _la ).

  FIG. 14 is a flowchart showing a procedure for setting the target valve closing timing and the target mechanical compression ratio of the intake valve 7 in the present embodiment.

As shown in FIG. 14, first, in step S11, the required load is detected by the load sensor 41. Next, in step S12, the target valve closing timing of the intake valve 7 is calculated based on the required load detected in step S11. Specifically, as shown in FIG. 9, when the engine load is equal to or greater than the engine load L ₁ corresponding to the limit valve closing timing, the target valve closing timing of the intake valve 7 is advanced as the engine load increases. When it is ₁ or less, the target valve closing timing of the intake valve 7 is the limit valve closing timing.

Next, in step S13, the engine speed is detected based on the output of the crank angle sensor 42, and the engine speed is corrected using a map as shown in FIG. 13A based on the detected engine speed. A coefficient k _re is calculated. In step S14, the closing timing of the intake valve 7 is calculated based on a command value from the ECU 30 to the variable valve timing mechanism B, and based on the calculated closing timing of the intake valve 7, FIG. The valve closing timing correction coefficient k _ivc is calculated using the map as shown. Thereafter, in step S15, the lift amount during the valve opening period of the intake valve 7 is calculated based on the command value from the ECU 30 to the variable valve timing mechanism B, and based on the calculated lift amount, FIG. The lift amount correction coefficient k _la is calculated using the map as shown in FIG.

Next, in step S16, the reference deviation amount ΔV _base is added to the rotation speed correction coefficient k _re calculated in step S13, the valve closing timing correction coefficient k _ivc calculated in step S14, and the lift amount calculated in step S15. The deviation amount ΔV is calculated by multiplying the correction coefficient k _la (ΔV = ΔV _base · k _re · _kiv · k _la ).

  In step S17, the combustion chamber volume at this timing is calculated based on the closing timing of the intake valve 7 calculated in step S14, and is calculated in step S16 from the calculated combustion chamber volume at the intake valve 7 closing timing. The compression start volume is calculated by subtracting the deviation amount ΔV.

  In step S18, the mechanical compression ratio is calculated based on the compression start volume calculated in step S17 and the target actual compression ratio. That is, the target mechanical compression ratio is set so that the combustion chamber volume when the piston is at the compression top dead center is a value obtained by dividing the compression start volume by the target actual compression ratio.

  In the above embodiment, the target mechanical compression ratio is set based on the calculated compression start volume. However, other control parameters may be set based on the calculated compression start volume.

  For example, if the temperature and pressure of the intake gas in the combustion chamber 5 when the intake valve is closed can be detected or calculated, the temperature and pressure of the intake gas and the compression start volume in the combustion chamber 5 when the intake valve is closed The compression end temperature and the compression end pressure can be calculated based on the combustion chamber volume when the piston is at the compression top dead center.

  Therefore, based on the compression end temperature and the compression end pressure calculated in this way, the ignition timing by the spark plug 6 and, when the fuel injection valve 13 is disposed in the combustion chamber 5, from the fuel injection valve 13. The fuel injection timing may be set. In this case, for example, the ignition timing by the spark plug 6 is retarded as the compression end temperature increases.

DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 3 Cylinder head 4 Piston 5 Combustion chamber 7 Intake valve 16 Actuator 17 Throttle valve A Variable compression ratio mechanism B Variable valve timing mechanism

Claims (8)

In a control device for an internal combustion engine comprising a variable valve timing mechanism capable of changing a closing timing of an intake valve, which is a timing at which the lift amount of the intake valve becomes zero,
Based on engine operating parameters including at least one of the engine speed, intake valve closing timing, and intake valve lift during the intake valve opening period, the combustion chamber volume and piston at the closing timing of the intake valve A deviation amount calculating means for calculating a deviation amount from the combustion chamber volume at the compression action start time at which the actual compression action starts,
A valve closing time volume calculating means for calculating the combustion chamber volume at the actual closing timing of the intake valve;
The combustion chamber volume at the compression action start timing is calculated by subtracting the deviation amount calculated by the deviation amount calculation means from the combustion chamber volume at the actual valve closing timing of the intake valve calculated by the valve closing volume calculation means. Compression start volume calculating means;
Control parameter setting means for setting a control parameter of the internal combustion engine based on the combustion chamber volume at the compression action start time calculated by the compression start volume calculation means ,
The control device for an internal combustion engine, wherein the divergence amount calculating means calculates the divergence amount by correcting a predetermined reference divergence amount based on a closing timing of the intake valve . The deviation amount calculating means, the reference deviation amount to a predetermined further the deviation by correcting based on at least one of the lift amount of the intake valve in the engine speed及 Beauty intake valve during the valve open period The control device for an internal combustion engine according to claim 1, wherein the amount is calculated.   The above reference deviation amount assumes that the compression of the intake gas in the combustion chamber after the intake valve is closed in a predetermined engine operating state is adiabatic compression, and the adiabatic compression is also performed before the intake valve is closed. The combustion chamber volume at the time when the pressure in the combustion chamber becomes a predetermined pressure lower than the pressure in the combustion chamber at the intake valve closing timing, and the combustion chamber volume at the closing timing of the intake valve The control device for an internal combustion engine according to claim 2, wherein the control device is a difference.   The control apparatus for an internal combustion engine according to claim 3, wherein the predetermined pressure is an intake pipe pressure in the predetermined engine operating state.   The control device for an internal combustion engine according to any one of claims 2 to 4, wherein a correction amount with respect to a reference deviation amount when calculating the deviation amount is increased as the engine speed increases.   The internal combustion engine according to any one of claims 2 to 5, wherein a correction amount with respect to a reference divergence amount when calculating the divergence amount is increased as a lift amount of the intake valve during an intake valve opening period is decreased. Control device. The control device for an internal combustion engine according to any one of claims 1 to 6 , further comprising a variable compression ratio mechanism capable of changing a mechanical compression ratio, wherein the control parameter of the internal combustion engine is a mechanical compression ratio. The control device for an internal combustion engine according to any one of claims 1 to 6 , wherein the control parameter of the internal combustion engine is at least one of fuel injection timing from a fuel injection valve and ignition timing by a spark plug.

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