JP2012233439A - Spark-ignition internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To enable a fuel consumption to surely be improved with a simpler configuration of an internal combustion engine including a variable compression ratio mechanism that can change a mechanical compression ratio of the internal combustion engine even when an engine load is within an extremely low load range.SOLUTION: The spark-ignition internal combustion engine includes the variable compression ratio mechanism that can change the mechanical compression ratio, and a variable valve timing mechanism that can control a closing timing of an intake valve. The mechanical compression ratio during a low engine load operation is increased, compared to during an engine mid-high load operation. However, even during the low engine load operation, if the engine load is in the extremely low load range, the mechanical compression ratio is reduced, compared to during the low engine load operation.

Description

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

  For the purpose of improving the fuel consumption performance and output performance of the internal combustion engine, a variable compression ratio mechanism capable of changing the mechanical compression ratio of the internal combustion engine and a variable valve timing mechanism capable of controlling the closing timing of the intake valve are provided. A spark ignition internal combustion engine has been proposed. For example, 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 are provided, and the actual compression ratio is kept constant during engine load operation and engine load operation. The mechanical compression ratio is increased and the closing timing of the intake valve is delayed as the engine load is lowered in the held state, and the maximum expansion ratio is obtained during engine low load operation in order to improve thermal efficiency. There is known a spark ignition internal combustion engine configured to maximize the mechanical compression ratio.

  A variable compression ratio mechanism that can change the mechanical compression ratio uses, for example, a cam mechanism to move a cylinder block that is an engine element constituting a combustion chamber of an internal combustion engine relative to a crankcase. By changing the volume of the combustion chamber, the mechanical compression ratio of the internal combustion engine is changed, and when the mechanical compression ratio is lowered, the combustion chamber volume is controlled to be increased. When the mechanical compression ratio is increased, the combustion chamber volume is controlled to be reduced.

JP 2005-226572 A

  In the variable compression ratio mechanism as described above, when the mechanical compression ratio is increased, the ratio of the surface area of the combustion chamber to the volume of the combustion chamber (hereinafter referred to as S / V ratio) increases. The phenomenon that the heat generated in the combustion chamber is easily dissipated compared to when the S / V ratio is small occurs, which may increase the cooling loss and reduce the effect of improving the thermal efficiency.

  Therefore, in an internal combustion engine configured to increase the mechanical compression ratio so that a high expansion ratio can be obtained during engine low load operation in order to improve thermal efficiency, the engine load is in an extremely low load region However, when a higher compression ratio of the mechanical compression ratio is executed, the effect of increasing the cooling loss due to the higher compression ratio of the mechanical compression ratio is larger than the effect of improving the thermal efficiency by increasing the expansion ratio. It may cause a situation of deteriorating.

  In this regard, in Patent Document 1, in a variable compression ratio internal combustion engine in which the mechanical compression ratio is changed by changing the volume of the combustion chamber, the cooling capacity for cooling the engine is higher at the time of the high compression ratio than at the time of the low compression ratio. In other words, as the S / V ratio increases and the heat generated in the combustion chamber becomes easier to dissipate, cooling of the engine by the cooling means is suppressed, and as a result, the amount of heat released from the combustion chamber is suppressed. Is disclosed.

  However, the variable compression ratio internal combustion engine disclosed in Patent Document 1 is configured to reduce the pumping amount of cooling water in order to suppress cooling of the engine. In such a configuration, the mechanical compression ratio is reduced. In addition to the above control, a configuration that controls the pumping amount of the cooling water according to the size of the mechanical compression ratio is required, and after the pumping amount of the cooling water is reduced, the cooling loss is reduced. It is possible that a considerable time delay may occur.

  In view of the above problems, the present invention provides a spark ignition type internal combustion engine including a variable compression ratio mechanism capable of changing the mechanical compression ratio of the internal combustion engine and a variable valve timing mechanism capable of controlling the closing timing of the intake valve. However, a simpler configuration is possible even when the engine load is in the extremely low load region without requiring a configuration for controlling the pumping amount of the cooling water according to the size of the mechanical compression ratio. It is another object of the present invention to provide a spark ignition type internal combustion engine that can improve fuel consumption reliably.

  According to the first aspect of the present invention, the variable compression ratio mechanism capable of changing the mechanical compression ratio and the variable valve timing mechanism capable of controlling the closing timing of the intake valve are provided. The mechanical compression ratio is higher than that during load operation, and the mechanical compression ratio is lower than that during low engine load operation when the engine load is in the extremely low load range even during low engine load operation. A spark ignition internal combustion engine is provided.

  That is, according to the first aspect of the present invention, the mechanical compression ratio is made higher in the engine low load operation than in the engine middle high load operation in order to improve the thermal efficiency. In an extremely low load region where both the engine speed and the engine speed do not reach the predetermined level, the mechanical compression ratio is reduced as compared with the engine low load operation. According to the present invention having such a configuration, when the engine load is in an extremely low load region, the S / V ratio is reduced and the cooling loss is improved by reducing the mechanical compression ratio. This makes it possible to improve fuel consumption. In addition, regarding the descriptions of “engine low load” and “engine medium high load”, the region that the internal combustion engine can take is divided into three regions, which are divided into “low load region”, “medium load region” in order from the lower side, “Low load range” when “high load range” is set corresponds to “engine low load”, and a combination of “medium load range” and “high load range” corresponds to “engine medium high load”. In addition, as described above, the description of “extremely low load region” corresponds to a region where both the engine load and the engine speed have not reached a predetermined level in the low load region. For example, it corresponds to a region where the average effective pressure (Pme) of the engine is smaller than 0.2 MPa and the engine speed is smaller than 1200 rpm.

  According to the second aspect of the present invention, the mechanical compression ratio is set to the maximum mechanical compression ratio so that the maximum expansion ratio is obtained at the time of engine low load operation, and the engine load is extremely low even at the time of engine low load operation. The spark ignition type internal combustion engine according to claim 1, wherein the mechanical compression ratio is reduced from the maximum mechanical compression ratio when in the region.

  According to a third aspect of the present invention, the closing timing of the intake valve moves away from the intake bottom dead center until the limit closing timing at which the amount of intake air supplied into the combustion chamber can be controlled as the engine load decreases. A spark ignition internal combustion engine according to claim 2 is provided which is moved in a direction.

  According to the invention described in each claim, the variable compression ratio mechanism capable of changing the mechanical compression ratio and the variable valve timing mechanism capable of controlling the closing timing of the intake valve are provided. In a spark ignition type internal combustion engine in which the mechanical compression ratio is higher than that at the time of load operation, even when the engine load is in an extremely low load region, the fuel consumption should be improved with a simpler structure and surely. It has the common effect of making it possible.

1 is an overall view of an internal combustion engine. It is a disassembled perspective view of a variable compression ratio mechanism. 1 is a schematic side sectional view of an internal combustion engine. It is a figure which shows a variable valve timing mechanism. It is a figure which shows the lift amount of an intake valve and an exhaust valve. It is a figure for demonstrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. It is a figure which shows the relationship between theoretical thermal efficiency and an expansion ratio. It is a figure for demonstrating a normal cycle and a super-high expansion ratio cycle. It is a figure which shows one Embodiment of the map used for control of the mechanical compression ratio in the past. It is a figure which shows an example of the tendency of the sensitivity of a cooling loss in case the mechanical compression ratio control by the map shown in FIG. 9 is made. It is a figure which shows an example of the tendency of a fuel consumption improvement in case mechanical compression ratio control is made by the map shown in FIG. It is a figure which shows an example of the map used for control of the mechanical compression ratio by this invention.

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

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

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

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

  The electronic control unit 30 is composed of a digital computer, and is connected to each other by a bidirectional bus 31. A ROM (read only memory) 32, a RAM (random access memory) 33, a CPU (microprocessor) 34, an input port 35 and an output port 36. It comprises. Output signals of the intake air amount detector 18, the air-fuel ratio sensor 21, the relative position sensor 22, the valve timing sensor 23, and the throttle opening sensor 24 are input to the input port 35 via the corresponding AD converters 37. A load sensor 41 that generates an output voltage proportional to the depression amount L of the accelerator pedal 40 is connected to the accelerator pedal 40, and the output voltage of the load sensor 41 is input to the input port 35 via the corresponding AD converter 37. Is done. Further, a crank angle sensor 42 that generates an output pulse every time the crankshaft rotates, for example, 30 ° is connected to the input port 35. On the other hand, the output port 36 is connected to the spark plug 6, the fuel injection valve 13, the throttle valve driving actuator 16, the variable compression ratio mechanism A, and the variable valve timing mechanism B through corresponding drive circuits 38.

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

  As shown in FIG. 2, a pair of camshafts 54 and 55 are provided, and on each camshaft 54 and 55, a circular cam 58 is rotatably inserted into each cam insertion hole 53. It is fixed. These circular cams 58 are coaxial with the rotational axes of the camshafts 54 and 55. On the other hand, on both sides of each circular cam 58, as shown in FIG. 3, an eccentric shaft 57 eccentrically arranged with respect to the rotation axis of each camshaft 54, 55 extends. 56 is mounted eccentrically and rotatable. As shown in FIG. 2, these circular cams 56 are arranged on both sides of each circular cam 58, and these circular cams 56 are rotatably inserted into the corresponding cam insertion holes 51. As shown in FIG. 2, a cam rotation angle sensor 25 that generates an output signal representing the rotation angle of the camshaft 55 is attached to the camshaft 55.

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

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

  3A to 3C, the relative positions of the crankcase 1 and the cylinder block 2 are determined by the distance between the center a of the circular cam 58 and the center c of the circular cam 56. As the distance between the center a of 58 and the center c of the circular cam 56 increases, the cylinder block 2 moves away from the crankcase 1. That is, the variable compression ratio mechanism A changes the relative position between the crankcase 1 and the cylinder block 2 by a crank mechanism using a rotating cam. When the cylinder block 2 moves away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is positioned at the compression top dead center. Therefore, by rotating the camshafts 54 and 55, the piston 4 is compressed at the top dead center. The volume of the combustion chamber 5 when it is located at can be changed.

  As shown in FIG. 2, in order to rotate the camshafts 54 and 55 in opposite directions, a pair of worms 61 and 62 having opposite spiral directions are attached to the rotation shaft of the drive motor 59, respectively. Worm wheels 63 and 64 that mesh with the worms 61 and 62 are fixed to the ends of the camshafts 54 and 55, respectively. In this embodiment, by driving the drive motor 59, the volume of the combustion chamber 5 when the piston 4 is located at the compression top dead center can be changed over a wide range.

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

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

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

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

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

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

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

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

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

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

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

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

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

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

  On the other hand, under such circumstances, it is considered to increase the theoretical thermal efficiency while strictly dividing the mechanical compression ratio and the actual compression ratio. As a result, the theoretical thermal efficiency is governed by the expansion ratio, and the actual compression ratio is compared to the theoretical thermal efficiency. Was found to have little effect. That is, if the actual compression ratio is increased, the explosive force is increased, but a large amount of energy is required for compression. Thus, even if the actual compression ratio is increased, the theoretical thermal efficiency is hardly increased.

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

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

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

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

  Next, an embodiment of overall operation control in the spark ignition internal combustion engine of the present invention will be schematically described. As described above, the normal cycle shown in FIG. 8A is executed during engine high load operation. Therefore, since the mechanical compression ratio is lowered, the expansion ratio is low, and the closing timing of the intake valve 7 is advanced as shown by the solid line in FIG. 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, when the engine load becomes low, the closing timing of the intake valve 7 is delayed to reduce the intake air amount. Further, at this time, the mechanical compression ratio is increased as the engine load is lowered so that the actual compression ratio is kept substantially constant. Therefore, the expansion ratio is also increased as the engine load is lowered. At this time, the throttle valve 17 is kept fully open, and therefore the amount of intake air supplied into the combustion chamber 5 is controlled by changing the closing timing of the intake valve 7 regardless of the throttle valve 17. Has been.

  As described above, when the engine load is reduced from the engine high load operation state, the mechanical compression ratio is increased as the intake air amount is decreased while the actual compression ratio is substantially constant. That is, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is decreased in proportion to the decrease in the intake air amount. Therefore, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the intake air amount.

  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 L1 slightly close to the low load, the mechanical compression ratio becomes a limit mechanical compression ratio (upper limit mechanical compression) that becomes the structural limit of the combustion chamber 5 Ratio). When the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is held at the limit mechanical compression ratio in a region where the load is lower than the engine load L1 when the mechanical compression ratio reaches the limit mechanical compression ratio. Accordingly, the mechanical compression ratio is maximized and the expansion ratio is maximized at the time of low engine load operation and low engine load operation, that is, at the engine low load operation side. In other words, the mechanical compression ratio is maximized so that the maximum expansion ratio is obtained on the engine low load operation side.

  On the other hand, when the engine load decreases to L1, the closing timing of the intake valve 7 becomes the limit closing timing at which the amount of intake air supplied into the combustion chamber 5 can be controlled. When the closing timing of the intake valve 7 reaches the limit closing timing, the closing timing of the intake valve 7 is reduced in a region where the load is lower than the engine load L1 when the closing timing of the intake valve 7 reaches the closing timing. It is held at the limit closing timing.

  When the closing timing of the intake valve 7 is held at the limit closing timing, the amount of intake air can no longer be controlled by changing the closing timing of the intake valve 7. At this time, that is, in the region where the load is lower than the engine load L1 when the closing timing of the intake valve 7 reaches the limit closing timing, the amount of intake air supplied into the combustion chamber 5 is controlled by the throttle valve 17, As the engine load becomes lower, the opening degree of the throttle valve 17 is made smaller.

  On the other hand, 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. Therefore, in the embodiment according to the present invention, the closing timing of the intake valve 7 moves away from the intake bottom dead center BDC until the limit closing timing L1 that can control the amount of intake air supplied into the combustion chamber as the engine load decreases. It will be moved in the direction.

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

  By the way, by using a cam mechanism or the like to move the cylinder block, which is an engine element constituting the combustion chamber of the internal combustion engine, relative to the crankcase, the volume of the combustion chamber is changed and the mechanical compression ratio is increased. In the variable compression ratio mechanism configured to be changed, when the mechanical compression ratio is increased, the ratio of the surface area of the combustion chamber to the volume of the combustion chamber (hereinafter referred to as S / V ratio) is increased. Thus, the phenomenon that the heat generated in the combustion chamber becomes easier to dissipate than when the S / V ratio is small may occur, thereby increasing the cooling loss and reducing the effect of improving the thermal efficiency.

  Therefore, in an internal combustion engine configured to increase the mechanical compression ratio so that a high expansion ratio can be obtained during engine low load operation in order to improve thermal efficiency, the engine load is in an extremely low load region However, when a higher compression ratio of the mechanical compression ratio is executed, the effect of increasing the cooling loss due to the higher compression ratio of the mechanical compression ratio is larger than the effect of improving the thermal efficiency by increasing the expansion ratio. It may cause a situation of deteriorating.

  FIG. 9 is a diagram showing an embodiment of a map used for controlling the mechanical compression ratio in the related art. As shown in FIG. 9, the conventional control of the mechanical compression ratio is generally controlled such that the mechanical compression ratio is increased as the engine load decreases. FIG. 10 shows an example of the tendency of cooling loss susceptibility when the mechanical compression ratio is controlled by the map shown in FIG. FIG. 11 shows an example of the tendency of the improvement in fuel efficiency when the mechanical compression ratio is controlled by the map shown in FIG.

  Referring to FIG. 10, it is understood that the cooling loss is large in the region A in which both the engine load and the engine speed do not reach the predetermined level, that is, in the extremely low load region even in the low load region. sell. For this reason, in the extremely low load region A, the degree of adverse effect due to cooling loss is greater than the effect of improving the thermal efficiency by increasing the mechanical compression ratio and increasing the expansion ratio. As shown, it can be understood that the degree of improvement in fuel consumption is small in the extremely low load region A, although the degree of improvement in fuel consumption is large in the low load region other than the extremely low load region. That is, from the viewpoint of improving fuel efficiency in the extremely low load region A, the effect of improving the cooling loss is more important than improving the thermal efficiency by increasing the mechanical compression ratio and increasing the expansion ratio.

  Based on this, in the present invention, in order to improve the thermal efficiency, the mechanical compression ratio is increased at the time of engine low load operation compared to the time of engine medium and high load operation, while the engine is in the low load region. In the extremely low load region A where both the load and the engine speed have not reached the predetermined level, the mechanical compression ratio is configured to be lower than that during engine low load operation. Note that the extremely low load region A in the present invention corresponds to a region in which both the engine load and the engine speed do not reach a predetermined level in the low load region as described above. . The extremely low load region in the present embodiment corresponds to a region where the average effective pressure (Pme) of the engine is smaller than 0.2 MPa and the engine speed is smaller than 1200 rpm, for example. However, the thresholds for the engine load and the engine speed for specifying the extremely low load region are not limited to this, and may be appropriately determined according to the engine specifications and the like.

  According to the present invention having such a configuration, when the engine load is in an extremely low load region, the S / V ratio is reduced and the cooling loss is improved by reducing the mechanical compression ratio. This makes it possible to improve fuel efficiency. An example of a map used for controlling the mechanical compression ratio according to the present invention is shown in FIG. In the present invention, when the engine load is in the extremely low load region shown as region A in FIG. 12, the mechanical compression ratio is reduced as compared with the engine low load operation.

  Incidentally, as a means for improving the cooling loss, a measure for improving the cooling loss by reducing the pumping amount of the engine cooling water according to the magnitude of the mechanical compression ratio can be considered, but according to the present invention, the mechanical compression ratio is reduced. Even when the engine load is in an extremely low load region without requiring a configuration for controlling the pumping amount of the cooling water according to the size, the fuel consumption can be reliably improved with a simpler configuration. Can make it possible.

  When the present invention is applied to a spark ignition internal combustion engine configured with a swirl control valve (SCV), when the engine load is in an extremely low load region, As compared with the mechanical compression ratio, the cooling loss in the extremely low load region may be improved by appropriately controlling the combustion speed in the extremely low load region by using the swirl control valve. For example, if the swirl flow formed in the cylinder becomes stronger, the flow velocity increases, so the amount of heat released from the wall surface in contact with the swirl flow, such as the inner peripheral surface of the cylinder or the top surface of the piston, increases and cooling loss increases. When the engine load is in the extremely low load range, control to stop the generation of the swirl flow by the swirl control valve or the flow rate of the swirl flow It may be configured such that control is performed so as to reduce.

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

Claims (3)

  Equipped with a variable compression ratio mechanism that can change the mechanical compression ratio and a variable valve timing mechanism that can control the closing timing of the intake valve, the mechanical compression ratio is higher when the engine is under low load operation than when during medium and high load operation A spark-ignition internal combustion engine that is compared and has a mechanical compression ratio that is lower than that during engine low-load operation when the engine load is in an extremely low load region even during engine low-load operation.   The mechanical compression ratio is set to the maximum mechanical compression ratio so that the maximum expansion ratio can be obtained during engine low-load operation. When the engine load is within the extremely low load region even during engine low-load operation, the mechanical compression ratio is the maximum machine compression ratio. The spark ignition internal combustion engine according to claim 1, wherein the compression ratio is reduced to a low compression ratio.   The closing timing of the intake valve is moved in a direction away from the intake bottom dead center until a limit closing timing capable of controlling the amount of intake air supplied into the combustion chamber as the engine load decreases. Spark ignition internal combustion engine.

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