Classifications

• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
  • F02B75/00—Other engines
  • F02B75/04—Engines with variable distances between pistons at top dead-centre positions and cylinder heads
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
  • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
  • F01L13/00—Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations
  • F01L13/0015—Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations for optimising engine performances by modifying valve lift according to various working parameters, e.g. rotational speed, load, torque
  • F01L13/0063—Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations for optimising engine performances by modifying valve lift according to various working parameters, e.g. rotational speed, load, torque by modification of cam contact point by displacing an intermediate lever or wedge-shaped intermediate element, e.g. Tourtelot
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
  • F02B75/00—Other engines
  • F02B75/04—Engines with variable distances between pistons at top dead-centre positions and cylinder heads
  • F02B75/041—Engines with variable distances between pistons at top dead-centre positions and cylinder heads by means of cylinder or cylinderhead positioning
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02D—CONTROLLING COMBUSTION ENGINES
  • F02D13/00—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
  • F02D13/02—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02D—CONTROLLING COMBUSTION ENGINES
  • F02D13/00—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
  • F02D13/02—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
  • F02D13/0223—Variable control of the intake valves only
  • F02D13/0226—Variable control of the intake valves only changing valve lift or valve lift and timing
  • F02D13/023—Variable control of the intake valves only changing valve lift or valve lift and timing the change of valve timing is caused by the change in valve lift, i.e. both valve lift and timing are functionally related
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02D—CONTROLLING COMBUSTION ENGINES
  • F02D13/00—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
  • F02D13/02—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
  • F02D13/0269—Controlling the valves to perform a Miller-Atkinson cycle
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02D—CONTROLLING COMBUSTION ENGINES
  • F02D15/00—Varying compression ratio
  • F02D15/04—Varying compression ratio by alteration of volume of compression space without changing piston stroke
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02D—CONTROLLING COMBUSTION ENGINES
  • F02D41/00—Electrical control of supply of combustible mixture or its constituents
  • F02D41/0002—Controlling intake air
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
  • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
  • F01L1/00—Valve-gear or valve arrangements, e.g. lift-valve gear
  • F01L1/34—Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift
  • F01L1/344—Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear
  • F01L1/3442—Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear using hydraulic chambers with variable volume to transmit the rotating force
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
  • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
  • F01L1/00—Valve-gear or valve arrangements, e.g. lift-valve gear
  • F01L1/34—Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift
  • F01L1/344—Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear
  • F01L1/3442—Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear using hydraulic chambers with variable volume to transmit the rotating force
  • F01L2001/34423—Details relating to the hydraulic feeding circuit
  • F01L2001/34426—Oil control valves
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02D—CONTROLLING COMBUSTION ENGINES
  • F02D13/00—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
  • F02D13/02—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
  • F02D13/0223—Variable control of the intake valves only
  • F02D13/0234—Variable control of the intake valves only changing the valve timing only
  • F02D13/0238—Variable control of the intake valves only changing the valve timing only by shifting the phase, i.e. the opening periods of the valves are constant
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02D—CONTROLLING COMBUSTION ENGINES
  • F02D41/00—Electrical control of supply of combustible mixture or its constituents
  • F02D41/0002—Controlling intake air
  • F02D2041/001—Controlling intake air for engines with variable valve actuation
• • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
  • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
  • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
  • Y02T10/00—Road transport of goods or passengers
  • Y02T10/10—Internal combustion engine [ICE] based vehicles
  • Y02T10/12—Improving ICE efficiencies
• • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
  • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
  • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
  • Y02T10/00—Road transport of goods or passengers
  • Y02T10/10—Internal combustion engine [ICE] based vehicles
  • Y02T10/40—Engine management systems

Definitions

• the present invention relates to a spark ignition type internal combustion engine.
• BACKGROUND ART Known in the art is a spark ignition type internal combustion engine provided with a variable compression ratio mechanism able to change a mechanical compression ratio and a variable valve timing mechanism able to control a closing timing of an intake valve, performing a supercharging action by a supercharger at the time of engine medium load operation and engine high load operation and, in the state holding an actual compression ratio fixed at the time of engine medium and high load operation, increasing the mechanical compression ratio and retarding the closing timing of the intake valve as the engine load becomes lower (for example, see Japanese Patent Publication (A) No. 2004-218522).
• the larger the expansion ratio the longer the period in an expansion stroke where a downward force acts on the piston, therefore the larger the expansion ratio, the more the heat efficiency is improved. Therefore, to raise the heat efficiency at the time of engine operation, it is preferable to make the mechanical compression ratio as high as possible and make the expansion ratio a large one.
• an object of the present invention is to provide a spark ignition type internal combustion engine able to maintain an exhaust purification catalyst at a relatively high temperature even when operating the internal combustion engine in the state of a large expansion ratio.
• the present invention provides a spark ignition type internal combustion engine described in the claims of the claim section as means for realizing the above object.
• variable compression ratio mechanism able to change a mechanical compression ratio
• an actual compression action start timing changing mechanism able to change a start timing of an actual compression action
• an exhaust valve wherein at the time of engine low load operation the mechanical compression ratio is maximized to obtain a maximum expansion ratio and the actual compression ratio is set so that no knocking occurs, wherein the maximum expansion ratio is 20 or more, and wherein the closing timing of the exhaust valve at the time of engine low load operation is made substantially intake top dead center.
• a spark ignition type internal combustion engine comprising a variable compression ratio mechanism able to change a mechanical compression ratio, an actual compression action start timing changing mechanism able to change a start timing of an actual compression action, and an exhaust variable valve timing mechanism able to change the closing timing of the exhaust valve, wherein at the time of engine low load operation the mechanical compression ratio is maximized to obtain a maximum expansion ratio and the actual compression ratio is set so that no knocking occurs, wherein the maximum expansion ratio is 20 or more, and wherein a settable region of the closing timing of the exhaust valve at the time of engine low load operation is limited more to an intake top dead center side than that at the time of engine high load operation.
• the closing timing of the exhaust valve is made substantially intake top dead center.
• the engine further comprises an intake variable valve timing mechanism able to change the opening timing of the intake valve, and the closing timing of the exhaust valve and the opening timing of the intake valve are controlled so that at the time of engine low load operation a period where the opening of the intake valve and the opening of the exhaust valve overlap is made minimum.
• the engine further comprises an intake variable valve timing mechanism able to change the opening timing of the intake valve, and the .closing timing of the exhaust valve and the opening timing of the intake valve are controlled so that at the time of engine low load operation the period where the opening of the intake valve and the opening of the exhaust valve overlap, becomes zero.
• the engine further comprises an intake valve opening timing changing mechanism able to change the opening timing of the intake valve and, and at the time of engine low load operation, the opening timing of the intake valve is made substantially intake top dead center.
• the actual compression ratio at the time of engine low load operation is made substantially the same as the actual compression ratio at the time of engine medium and high load operation.
• the actual compression ratio falls within a range of 9 to 11.
• the actual compression action start timing changing mechanism is comprised of an intake variable valve timing mechanism able to change the closing timing of the intake valve.
• the amount of intake air fed into the combustion chamber is controlled by changing the closing timing of the intake valve .
• the closing timing of the intake valve is shifted as the engine load becomes lower to a direction away from intake bottom dead center until a limit closing timing enabling control of the amount of intake air fed into the combustion chamber.
• the amount of intake air fed into the combustion chamber is controlled without regard to a throttle valve arranged in an engine intake passage by changing the closing timing of the intake valve.
• the throttle valve in a region of a load higher than the engine load when the closing timing of the intake valve reaches the limit closing timing, the throttle valve is held at the fully opened state.
• a throttle valve arranged in an engine intake passage is used to control the amount of intake air fed into the combustion chamber.
• the closing timing of the intake valve is held at the limit closing timing.
• the mechanical compression ratio is increased as the engine load becomes lower to the limit mechanical compression ratio.
• the mechanical compression ratio in a region of a load lower than the engine load when the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is held at the limit mechanical compression ratio.
• the exhaust purification catalyst since as much exhaust gas as possible is discharged from the combustion chamber to the exhaust purification catalyst, even if operating the internal combustion engine in the state of a large expansion ratio, the exhaust purification catalyst can be maintained at a relatively high temperature.
• FIG. 1 is an overview of a spark ignition type internal combustion engine.
• FIG. 2 is a disassembled perspective view of a variable compression ratio mechanism.
• FIGS. 3A and 3B are side cross-sectional views of the illustrated internal combustion engine.
• FIG. 4 is a view of a variable valve timing mechanism.
• FIGS. 5A and 5B are views showing the amounts of lift of the intake valve and exhaust valve.
• FIGS. 6A, 6B and 6C are views for explaining the mechanical compression ratio, actual compression ratio, and expansion ratio.
• FIG. 7 is a view showing the relationship between the theoretical thermal efficiency and expansion ratio.
• FIGS. 8A and 8B are views for explaining a normal cycle and superhigh expansion ratio cycle.
• FIG. 9 is a view showing the change in mechanical compression ratio etc. in accordance with the engine load.
• FIG. 1OA, 1OB and 1OC are views showing the changes in lift of the intake valve and exhaust valve.
• FIG. 11 is a view showing a region in which a closing timing of the exhaust valve in accordance with the mechanical compression ratio can be set.
• FIGS. 12A and 12B are views showing the changes in lift of the intake valve and exhaust valve.
• FIG. 13 is a flowchart for operational control.
• FIGS. 14A, 14B and 14C are views showing the target actual compression ratio etc.
• FIGS. 15A and 15B are views showing a map of the closing timing of the exhaust valve etc.
• FIG. 1 shows a side cross-sectional view of a spark ignition type internal combustion engine.
• 1 indicates a crank case, 2 a cylinder block, 3 a cylinder head, 4 a piston, 5 a combustion chamber, 6 a spark plug arranged at the top center of the combustion chamber 5, 7 an intake valve, 8 an intake port, 9 an exhaust valve, and 10 an exhaust port.
• the intake port 8 is connected through an intake tube 11 to a surge tank 12, while each intake tube 11 is provided with a fuel injector 13 for injecting fuel toward a corresponding intake port 8.
• each fuel injector 13 may be arranged at each combustion chamber 5 instead of being attached to each intake tube 11.
• the surge tank 12 is connected via an intake duct 14 to an outlet of the compressor 15a of the exhaust turbocharger 15, while an inlet of the compressor 15a is connected through an intake air amount detector 16 using for example a heating wire to an air cleaner 17.
• the intake duct 14 is provided inside it with a throttle valve 19 driven by an actuator 18.
• the exhaust port 10 is connected through the exhaust manifold 20 to the inlet of the exhaust turbine 15b of the exhaust turbocharger 15, while an outlet of the exhaust turbine 15b is connected through an exhaust pipe 21 to a catalytic converter 22 housing an exhaust purification catalyst.
• the exhaust pipe 21 has an air-fuel ratio sensor 23 arranged in it.
• the connecting part of the crank case 1 and the cylinder block 2 is provided with a variable compression ratio mechanism A able to change the relative positions of the crank case 1 and cylinder block 2 in the cylinder axial direction so as to change the volume of the combustion chamber 5 when the piston 4 is positioned at compression top dead center.
• a variable compression ratio mechanism A able to change the relative positions of the crank case 1 and cylinder block 2 in the cylinder axial direction so as to change the volume of the combustion chamber 5 when the piston 4 is positioned at compression top dead center.
• an intake variable valve timing mechanism B able to control the closing timing of the intake valve 7 to change the start timing of the actual compression action, and able to individually control the opening timing of the intake valve 7.
• an exhaust variable valve timing mechanism C able to individually control the opening timing and closing timing of the exhaust valve 7.
• the electronic control unit 30 is comprised of a digital computer provided with components connected with each other through a bidirectional bus 31 such as a ROM (read only memory) 32, RAM (random access memory) 33, CPU (microprocessor) 34, input port 35, and output port 36.
• ROM read only memory
• RAM random access memory
• CPU microprocessor
• the output signal of the intake air amount detector 16 and the output signal of the air-fuel ratio sensor 23 are input through the corresponding AD converter 37 to the input port 35.
• an accelerator pedal 40 is connected to a load sensor 41 generating an output voltage proportional to the amount of depression of the accelerator pedal 40.
• the output voltage of the load sensor 41 is input through a corresponding AD converter 37 to the input port 35.
• the input port 35 is connected to a crank angle sensor 42 generating an output pulse every time the crankshaft rotates by for example 30°.
• the output port 36 is connected through the drive circuit 38 to the spark plug 6, fuel injector 13, throttle valve drive actuator 18, variable compression ratio mechanism A, and intake variable valve timing mechanism B.
• FIG. 2 is a disassembled perspective view of the variable compression ratio mechanism A shown in FIG. 1, while FIGS. 3A and 3B are side cross-sectional views of the illustrated internal combustion engine.
• FIG. 2 at the bottom of the two side walls of the cylinder block 2, a plurality of projecting parts 50 separated from each other by a certain distance are formed. Each projecting part 50 is formed with a circular cross-section cam insertion hole 51.
• the top surface of the crank case 1 is formed with a plurality of projecting parts 52 separated from each other by a certain distance and fitting between the corresponding projecting parts 50. These projecting parts 52 are also formed with circular cross-section cam insertion holes 53.
• a pair of cam shafts 54, 55 is provided.
• Each of the cam shafts 54, 55 has circular cams 56 fixed on it able to be rotatably inserted in the cam insertion holes 51 at every other position.
• These circular cams 56 are coaxial with the axes of rotation of the cam shafts 54, 55.
• eccentric shafts 57 arranged eccentrically with respect to the axes of rotation of the cam shafts 54, 55.
• Each eccentric shaft 57 has other circular cams 58 rotatably attached to it eccentrically. As shown in FIG. 2, these circular cams 58 are arranged between the circular cams 56.
• the relative positions of the crank case 1 and cylinder block 2 are determined by the distance between the centers of the circular cams 56 and the centers of the circular cams 58.
• FIG. 2 to make the cam shafts 54, 55 rotate in opposite directions, the shaft of a drive motor 59 is provided with a pair of worm gears 61, 62 with opposite thread directions. Gears 63, 64 engaging with these worm gears 61, 62 are fastened to ends of the cam shafts 54, 55.
• the drive motor 59 may be driven to change the volume of the combustion chamber 5 when the piston 4 is positioned at compression top dead center over a broad range.
• the variable compression ratio mechanism A shown from FIGS. 1 to 3 shows an example. Any type of variable compression ratio mechanism may be used.
• FIG. 4 shows an intake variable valve timing mechanism B attached to the cam shaft 70 for driving the intake valve 7 in FIG. 1. Referring to FIG.
• the intake variable valve timing mechanism B is comprised of a cam phase changer Bl attached to one end of the cam shaft 70 and changing the phase of the cam of the cam shaft 70 and a cam actuation angle changer B2 arranged between the cam shaft 70 and the valve lifter 24 of the intake valve 7 and changing the actuation angle (working angle) of the cams of the cam shaft 70 to different actuation angles for transmission to the intake valve 7.
• FIG. 4 is a side sectional view and plan view of the cam actuation angle changer B2.
• this cam phase changer Bl is provided with a timing pulley 71 made to rotate by an engine crank shaft through a timing belt in the arrow direction, a cylindrical housing 72 rotating together with the timing pulley 71, a shaft 73 able to rotate together with a cam shaft 70 and rotate relative to the cylindrical housing 72, a plurality of partitions 74 extending from an inside circumference of the cylindrical housing 72 to an outside circumference of the shaft 73, and vanes 75 extending between the partitions 74 from the outside circumference of the shaft 73 to the inside circumference of the cylindrical housing 72, the two sides of the vanes 75 formed with advancing use hydraulic chambers 76 and retarding use hydraulic chambers 77.
• the feed of working oil to the hydraulic chambers 76, 77 is controlled by a working oil feed control valve 78.
• This working oil feed control valve 78 is provided with hydraulic ports 79, 80 connected to the hydraulic chambers 76, 77, a feed port 82 for working oil discharged from a hydraulic pump 81, a pair of drain ports 83, 84, and a spool valve 85 for controlling connection and disconnection of the ports 79, 80, 82, 83, 84.
• the spool valve 85 is made to move to downward in FIG. 4, working oil fed from the feed port 82 is fed through the hydraulic port 79 to the advancing use hydraulic chambers 76, and working oil in the retarding use hydraulic chambers 77 is drained from the drain port 84.
• the shaft 73 is made to rotate relative to the cylindrical housing 72 in the arrow X- direction.
• the spool valve 85 is made to move upward in FIG. 4, working oil fed from the feed port 82 is fed through the hydraulic port 80 to the retarding use hydraulic chambers 77, and working oil in the advancing use hydraulic chambers 76 is drained from the drain port 83.
• the shaft 73 is made to rotate relative to the cylindrical housing 72 in the direction opposite to the arrows X.
• this cam actuation angle changer B2 is provided with a control rod 90 arranged in parallel with the cam shaft 70 and made to move by an actuator 91 in the axial direction, an intermediate cam 94 engaging with a cam 92 of the cam shaft 70 and slidably fitting with a spline 93 formed on the control rod 90 and extending in the axial direction, and a pivoting cam 96 engaging with a valve lifter 24 for driving the intake valve 7 and slidably fitting with a spline 95 extending in a spiral formed on the control rod 90.
• the pivoting cam 96 is formed with a cam 97.
• the cam 92 causes the intermediate cam 94 to pivot by exactly a constant angle at all times.
• the pivoting cam 96 is also made to pivot by exactly a constant angle.
• the intermediate cam 94 and pivoting cam 96 are supported not movably in the axial direction of the control rod 90, therefore when the control rod 90 is made to move by the actuator 91 in the axial direction, the pivoting cam 96 is made to rotate relative to the intermediate cam 94.
• the opening time period and amount of lift of the intake valve 7 become further smaller. That is, by using the actuator 91 to change the relative rotational position of the intermediate cam 94 and pivoting cam 96, the opening time period of the intake valve 7 can be freely changed. However, in this case, the amount of the lift of the intake valve 7 becomes smaller the shorter the opening time of the intake valve 7.
• the cam phase changer Bl can be used to freely change the opening timing of the intake valve 7 and the cam actuation angle changer B2 can be used to freely change the opening time period of the intake valve 7 in this way, so both the cam phase changer Bl and cam actuation angle changer B2, that is, the intake variable valve timing mechanism B, may be used to freely change the opening timing and opening time period of the intake valve 7, that is, the opening timing and closing timing of the intake valve 7.
• the intake variable valve timing mechanism B shown in FIGS. 1 and 4 shows an example. It is also possible to use various types of variable valve timing mechanisms other than the example shown in FIGS. 1 and 4. Further, the exhaust variable valve timing mechanism C also basically has a configuration similar to the intake variable valve timing mechanism B and can freely change the opening timing and opening time period of the exhaust valve 9, that is, the opening timing and closing timing of the exhaust valve 9.
• FIGS. 6A, 6B and 6C show for explanatory purposes an engine with a volume of the combustion chambers of 50 ml and a stroke volume of the piston of 500 ml.
• the combustion chamber volume shows the volume of the combustion chamber when the piston is at compression top dead center.
• FIG. 6A explains the mechanical compression ratio.
• the mechanical compression ratio is a value determined mechanically from the stroke volume of the piston and combustion chamber volume at the time of a compression stroke. This mechanical compression ratio is expressed by (combustion chamber volume+stroke volume) /combustion chamber volume. In the example shown in FIG. ⁇ A, this _ _
• FIG. 6B explains the actual compression ratio.
• This actual compression ratio is a value determined from the actual stroke volume of the piston from when the compression action is actually started to when the piston reaches top dead center and the combustion chamber volume.
• FIG. 6C explains the expansion ratio.
• FIG. 7 shows the relationship between the theoretical thermal efficiency and the expansion ratio
• FIGS. 8A and 8B show a comparison between the normal cycle and superhigh expansion ratio cycle used selectively in accordance with the load in the present invention
• FIG. 8A shows the normal cycle in which the intake valve closes near the bottom dead center and the compression action by the piston is started from near substantially compression bottom dead center.
• the combustion chamber volume is made 50 ml
• the stroke volume of the piston is made 500 ml.
• the actual compression ratio is also about 11
• the solid line in FIG. 7 shows the change in the theoretical thermal efficiency in the case where the actual compression ratio and expansion ratio are substantially equal, that is, in the normal cycle.
• the larger the expansion ratio that is, the higher the actual compression ratio, the higher the theoretical thermal efficiency. Therefore, in a normal cycle, to raise the theoretical thermal efficiency, the actual compression ratio should be made higher.
• the theoretical thermal efficiency cannot be made sufficiently high.
• the inventors strictly differentiated between the mechanical compression ratio and actual compression ratio and studied the theoretical thermal efficiency and as a result discovered that in the theoretical thermal efficiency, the expansion ratio is dominant, and the theoretical thermal efficiency is not affected much at all by the actual compression ratio. That is, if raising the actual compression ratio, the explosive force rises, but compression requires a large energy, accordingly even if raising the actual compression ratio, the theoretical thermal efficiency will not rise much at all.
• the broken line in FIG. 7 shows the theoretical thermal efficiency in the case of fixing the actual compression ratio at 10 and raising the expansion ratio in that state.
• FIG. 8B shows an example of the case when using the variable compression ratio mechanism A and variable valve timing mechanism B to maintain the actual compression ratio at a low value and raise the expansion ratio.
• variable compression ratio mechanism A is used to lower the combustion chamber volume from 50 ml to 20 ml.
• the intake variable valve timing mechanism B is used to retard the closing timing of the intake valve until the actual stroke volume of the piston changes from 500 ml to 200 ml.
• the actual compression ratio is about 11 and the expansion ratio is 11.
• FIG. 8A is set. This is the basic feature of the present invention.
• FIG. 9 shows the operational control as a whole at the time of steady operation at a low engine speed. Below, the operational control as a whole will be explained with reference to FIG. 9.
• FIG. 9 shows the changes in the mechanical compression ratio, expansion ratio, closing timing of the intake valve 7, actual compression ratio, the amount of intake air, opening degree of the throttle valve 17, and pumping loss, along with the engine load.
• the average air-fuel ratio in the combustion chamber 5 is feedback controlled to the stoichiometric air-fuel ratio based on the output signal of the air-fuel ratio sensor 23.
• the mechanical compression ratio is increased along with the fall in the amount of intake air under a substantially constant actual compression ratio. That is, the volume of the combustion chamber 5 when the piston 4 reaches compression top dead center is reduced proportionally to the reduction in the amount of intake air. Therefore the volume of the combustion chamber 5 when the piston 4 reaches compression top dead center changes proportionally to the amount of intake air. Note that at this time, the air-fuel ratio in the combustion chamber 5 becomes the stoichiometric air-fuel ratio, so the volume of the combustion chamber 5 when the piston 4 reaches compression top dead center changes proportionally to the amount of fuel.
• the mechanical compression ratio is further increased.
• the mechanical compression ratio reaches the limit mechanical compression ratio corresponding to the structural limit of the combustion chamber 5, in the region of a load lower than the engine load Ll when the mechanical compression ratio reaches the limit mechanical compression ratio, the mechanical compression ratio is held at the limit mechanical compression ratio. Therefore at the time of engine low load operation, the mechanical compression ratio becomes maximum, and the expansion ratio also becomes maximum.
• the mechanical compression ratio is made maximum. Further, at this time, the actual compression ratio is maintained at an actual compression ratio substantially the same as that at the time of engine medium and high load operation. On the other hand, as shown by the solid line in FIG.
• the closing timing of the intake valve 7 is retarded further to the limit closing timing enabling control of the amount of intake air fed to the combustion chamber 5 the more the engine load becomes lower.
• the closing timing of the intake valve 7 is held at the limit closing timing. If the closing timing of the intake valve 7 is held at the limit closing timing, the amount of intake air will no longer be able to be controlled by the change of the closing timing of the intake valve 7. Therefore, the amount of intake air has to be controlled by some other method.
• the throttle valve 17 is used to control the amount of intake air fed to the combustion chamber 5.
• the throttle valve 17 is used to control the amount of intake air, as shown in FIG. 9, the pumping loss increases. Note that to prevent such pumping loss from occurring, in the region of a load lower than the engine load L2 when the closing timing of the intake valve 7 reaches the limit closing timing, in the state holding the throttle valve 17 fully opened or substantially fully opened, the air-fuel ratio may be made larger the lower the engine load.
• the fuel injector 13 is preferably arranged in the combustion chamber 5 to perform stratified combustion.
• the actual compression ratio at this time is made the range of the actual compression ratio about at the time of engine medium and high load operation ⁇ 10 percent, preferably ⁇ 5 percent.
• the actual compression ratio at the time of engine low speed is made about 10+1, that is, from 9 to 11.
• the expansion ratio is made 26. The higher this expansion ratio, the better, but if 20 or more, a considerably high theoretical thermal efficiency can be obtained.
• variable compression ratio mechanism A is formed so that the expansion ratio becomes 20 or more.
• the mechanical compression ratio is changed continuously in accordance with the engine load.
• the mechanical compression ratio can also be changed in stages in accordance with the engine load.
• the harmful ingredients contained in exhaust gas are removed by providing inside the engine exhaust passage a three-way catalyst, NO x storing and reducing catalyst, or other exhaust purification catalyst.
• Such an exhaust purification catalyst cannot effectively remove the harmful ingredients in the exhaust gas unless its temperature becomes the activation temperature or more.
• the temperature of the exhaust gas is considerably higher than the activation temperature, so the exhaust gas is made to flow into the exhaust purification catalyst to maintain the temperature of the exhaust purification catalyst at the activation temperature or more.
• the superhigh expansion ratio cycle shown in FIG. 8B is executed, the temperature of the exhaust gas exhausted from the combustion chamber 5 to - 2 A -
• the exhaust manifold 20 will only become slightly higher than the activation temperature, so even if making the exhaust gas flow into the exhaust purification catalyst, it becomes difficult to maintain the temperature of the exhaust purification catalyst at the activation temperature or more. Therefore, when the superhigh expansion ratio cycle is executed, to maintain the temperature of the exhaust purification catalyst at the activation temperature or more, it is necessary to make as much exhaust gas as possible flow into the exhaust purification catalyst.
• FIGS. 1OA to 1OC let us consider the relationship between the closing timing of the exhaust valve 9 and the flow rate of exhaust gas exhausted from the combustion chamber 5 to the exhaust manifold 20.
• FIG. 1OA shows the changes in lift of the exhaust valve 9 and the intake valve 7 in the case where the exhaust valve 9 is closed at substantially intake top dead center
• FIG. 1OB shows the same in case where the exhaust valve 9 is closed before intake top dead center
• FIG. 1OC shows the same in the case where the exhaust valve 9 is closed after intake top dead center.
• the volume of the combustion chamber 5 when closing the exhaust valve 9 is larger than the volume of the combustion chamber when the piston is positioned at intake top dead center (combustion chamber volume) .
• exhaust gas corresponding to the volume of the combustion chamber 5 at the time of closing remains in the combustion chamber 5.
• a relatively large amount of exhaust gas remains in the combustion chamber 5. Therefore, it is not possible to sufficiently exhaust the exhaust gas in the combustion chamber 5 to the exhaust manifold 20 and the flow rate of exhaust gas into the exhaust purification catalyst becomes small.
• FIG. 1OB when closing the exhaust valve 9 before intake top dead center, the volume of the combustion chamber 5 when closing the exhaust valve 9 is larger than the volume of the combustion chamber when the piston is positioned at intake top dead center (combustion chamber volume) .
• the region where the closing timing of the exhaust valve 9 can be set is limited to the intake top dead center side.
• FIG. 11 is a view showing a region in which the closing timing of the exhaust valve 9 in accordance with the mechanical compression ratio can be set.
• the region in which the exhaust valve 9 can be set becomes the region between the settable maximum amount of advance and maximum amount of retardation.
• the amount of advance by which the closing timing of the exhaust valve 9 can be set is made smaller (later) the higher the mechanical compression ratio, while conversely the maximum amount of retardation by which the closing timing of the exhaust valve 9 can be set is made smaller (earlier) the higher the mechanical compression ratio.
• the region in which the closing timing of the exhaust valve 9 can be set becomes smaller the higher the mechanical compression ratio, that is, is more restricted the higher the mechanical compression ratio. For example, as shown in FIG.
• the region in which the closing timing of the exhaust valve 9 can be set is ⁇ TOC1
• the region in which the closing timing of the exhaust valve 9 can be set is made ⁇ TOC2 ( ⁇ TOC2 ⁇ TOC1) .
• the closing timing of the exhaust valve 9 may be made substantially intake top dead center.
• the exhaust valve 9 is made to close near intake top dead center, so as shown* in FIG. 1OB, compared with closing the exhaust valve 9 in advance of intake top dead center, the volume of the combustion chamber 5 at the time of closing of the exhaust valve 9 is small and therefore it is possible to reduce the amount of exhaust gas remaining in the combustion chamber 5 after closing the exhaust valve 9. Further, the exhaust valve 9 is made to close near intake top dead center, so as shown in FIG. 1OC, compared with when closing the exhaust valve 9 retarded from intake top dead center, the amount of exhaust gas flowing into the combustion chamber 5 in the exhaust gas flowing out into the exhaust port 10 can be reduced. For this reason, as shown in FIG. 1OA, when the exhaust valve 9 is made to close near intake top dead center, as shown in FIGS.
• the exhaust gas in the combustion chamber 5 can be sufficiently exhausted into the exhaust manifold 20 and the flow rate of the exhaust gas flowing into the exhaust purification catalyst can be increased. As a result, even at the time of low load operation where the superhigh expansion ratio cycle is executed, it is possible to maintain the exhaust purification catalyst at the activation temperature or more.
• substantially intake top dead center indicates within 10° before and after intake top dead center, preferably within 5° before and after intake top dead center. Further, if raising the mechanical compression ratio, the combustion chamber volume at intake top dead center becomes smaller and accordingly depending on the closing timing of the exhaust valve 9, the exhaust valve 9 will end up interfering with the piston 4.
• FIGS. 1OA to 1OC show the piston interference line showing the limit where the exhaust valve 9 or intake valve 7 interferes with the piston 4.
• the lift curve of the exhaust valve 9 interferes with the piston 4.
• the lift curve of the exhaust valve 9 intersects with the piston interference line. This means that when closing the exhaust valve 9 retarded from intake top dead center, while depending also on the extent of retardation, the exhaust valve 9 and piston 4 will end up interfering.
• the region in which the closing timing of the exhaust valve 9 can be set is limited to the intake top dead center side, in particular the amount of maximum retardation by which the closing timing of the exhaust valve 9 can be set is made smaller. For this reason, as shown in FIG. 1OA, even if the mechanical compression ratio becomes higher, the exhaust valve 9 can be prevented from interfering with the piston 4.
• FIGS. 12A and 12B consider the relationship between the overlap period where the opening time period of the intake valve 7 and the opening time period of the exhaust valve 9 overlap and the amount of exhaust gas exhausted from the combustion chamber 5 to the exhaust manifold 20.
• FIG. 12A shows the case where the overlap period is zero
• FIG. 12B shows the changes in lifts of the exhaust valve 9 and the intake valve 7 when the overlap period is large .
• the closing timing of the exhaust valve 7 and the opening timing of the intake valve 9 are controlled to become minimum in the range in which the overlap period can be set. Therefore, for example, in an internal combustion engine where the settable overlap period becomes 10° to 60°, when the mechanical compression ratio is high, the overlap period is made 10°, while in an internal combustion engine where the settable overlap period becomes 0° to 50°, when the mechanical compression ratio is high, the overlap period is made 0°.
• the overlap period when the mechanical compression ratio is high need not necessarily be made the minimum so long as it is shorter than the overlap period when the mechanical compression ratio is low. Therefore, for example, the overlap period when the mechanical compression ratio is high need only be 10° or less of the settable range even at the minimum. Further, as explained above, if raising the mechanical compression ratio, the combustion chamber volume at intake top dead center becomes smaller. Accordingly, depending on the opening timing of the intake valve 7, the intake valve 7 will end up interfering with the piston 4.
• FIGS. 12A and 12B show the piston interference line showing the limit where the exhaust valve 9 or intake valve 7 interferes with the piston 4. If the lift curve of the intake valve 7 intersects with the piston interference line, the intake valve 7 will interfere with the piston 4. Here, in FIG. 12B, the lift curve of the intake valve 7 intersects with the piston interference line. This means that if increasing the overlap period, the intake valve 7 and piston 4 will end up interfering with each other. That is, in the present embodiment, as explained above, the closing timing of the exhaust valve 9 is made substantially intake top dead center. The overlap period being large means that the opening timing of the intake valve 7 is made to greatly advance. If the opening timing of the intake valve 7 is made to greatly advance, the intake valve 7 and the piston 4 will end up interfering with each other.
• the overlap period is made the minimum, so the opening timing of the intake valve 7 is made substantially intake top dead center or less. For this reason, as shown in FIG. 12A, even if the mechanical compression ratio becomes high, the intake valve 7 can be prevented from interfering with the piston.
• FIG. 13 shows a control routine of operational control of a spark ignition type internal combustion engine of the present embodiment.
• the engine load L and engine speed Ne are fetched.
• the map shown in FIG. 14A is used to calculate the target actual compression ratio. As shown in FIG. 14A, this target actual compression ratio becomes higher the higher the engine speed Ne.
• the map shown in FIG. 14B is used to calculate the mechanical compression ratio CR. That is, the mechanical compression ratio CR required for making the actual compression ratio the target actual compression ratio is stored as a function of the engine load L and engine speed Ne in the form of a map as shown in FIG. 14B in advance in the ROM 32.
• This map is used to calculate the mechanical compression ratio CR. Further, the closing timing IC of the intake valve 7 required for feeding the required amount of intake air into the combustion chamber 5 is stored as a function of the engine load L and engine speed Ne in the form of a map as shown in FIG. 14C in advance in the ROM 32. At step 104, this map is used for calculating the closing timing IC of the intake valve 7.
• step 105 it is judged if the engine load L is smaller than a predetermined value L 3 .
• this predetermined value L 3 is, for example, made a value equal to the engine load at which when the engine load becomes smaller, the drop in the temperature of the exhaust gas may be accompanied with a drop in the temperature of the exhaust purification catalyst to below the activation temperature.
• the routine proceeds to step 106.
• the closing timing EC of the exhaust valve 9 is made substantially intake top dead center.
• the overlap period ⁇ OL is made the minimum and the routine proceeds to step 110.
• step 108 the map shown in FIG. 15A is used to calculate the closing timing EC of the exhaust valve 9
• step 109 the map shown in FIG. 15B is used to calculate the overlap period ⁇ OL. That is, the closing timing EC of the exhaust valve 9 and overlap period ⁇ OL are stored as functions of the engine load L and engine speed Ne in the form of the maps shown in FIGS. 15A and 15B in advance in the ROM 32. These maps are used to calculate the closing timing EC of the exhaust valve 9 and overlap period ⁇ OL.
• the routine proceeds to step 110.
• the mechanical compression ratio is made the mechanical compression ratio CR by controlling the variable compression ratio mechanism A, while the closing timing of the intake valve 7 is made the closing timing IC and the overlap period is made the overlap period ⁇ OL by controlling the intake variable valve timing mechanism B. Further, the closing timing of the exhaust valve 9 is made the closing timing EC by controlling the exhaust variable valve timing mechanism C.

Landscapes

• Engineering & Computer Science (AREA)
• Mechanical Engineering (AREA)
• General Engineering & Computer Science (AREA)
• Chemical & Material Sciences (AREA)
• Combustion & Propulsion (AREA)
• Output Control And Ontrol Of Special Type Engine (AREA)
• Electrical Control Of Air Or Fuel Supplied To Internal-Combustion Engine (AREA)
• Combined Controls Of Internal Combustion Engines (AREA)

Applications Claiming Priority (2)

Publications (1)

Family

ID=38325398

Family Applications (1)

Country Status (8)

Families Citing this family (12)

Family Cites Families (17)

• 2006
  • 2006-07-13 JP JP2006192832A patent/JP4259546B2/ja not_active Expired - Fee Related
• 2007
  • 2007-04-09 KR KR1020097000586A patent/KR101032288B1/ko not_active Expired - Fee Related
  • 2007-04-09 US US12/227,601 patent/US20090178632A1/en not_active Abandoned
  • 2007-04-09 CN CNA2007800240465A patent/CN101479453A/zh active Pending
  • 2007-04-09 WO PCT/JP2007/058219 patent/WO2008007488A1/en active Search and Examination
  • 2007-04-09 EP EP07741655A patent/EP2041411A1/en not_active Withdrawn
  • 2007-04-09 RU RU2009104935/06A patent/RU2411381C2/ru not_active IP Right Cessation
  • 2007-04-09 BR BRPI0714220-0A patent/BRPI0714220A2/pt not_active IP Right Cessation

Non-Patent Citations (1)

Also Published As

Similar Documents

Legal Events