JP4259546B2 - Spark ignition internal combustion engine - Google Patents

Spark ignition internal combustion engine Download PDF

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JP4259546B2
JP4259546B2 JP2006192832A JP2006192832A JP4259546B2 JP 4259546 B2 JP4259546 B2 JP 4259546B2 JP 2006192832 A JP2006192832 A JP 2006192832A JP 2006192832 A JP2006192832 A JP 2006192832A JP 4259546 B2 JP4259546 B2 JP 4259546B2
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Prior art keywords
valve
intake
compression ratio
closing timing
timing
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JP2008019799A (en
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大輔 秋久
大作 澤田
栄一 神山
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Toyota Motor Corp
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Toyota Motor Corp
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Priority to JP2006192832A priority Critical patent/JP4259546B2/en
Application filed by Toyota Motor Corp filed Critical Toyota Motor Corp
Priority to KR1020097000586A priority patent/KR101032288B1/en
Priority to PCT/JP2007/058219 priority patent/WO2008007488A1/en
Priority to US12/227,601 priority patent/US20090178632A1/en
Priority to RU2009104935/06A priority patent/RU2411381C2/en
Priority to EP07741655A priority patent/EP2041411A1/en
Priority to CNA2007800240465A priority patent/CN101479453A/en
Priority to BRPI0714220-0A priority patent/BRPI0714220A2/en
Publication of JP2008019799A publication Critical patent/JP2008019799A/en
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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00Other engines
    • F02B75/04Engines with variable distances between pistons at top dead-centre positions and cylinder heads
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L13/00Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations
    • F01L13/0015Modifications 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
    • F01L13/0063Modifications 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
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00Other engines
    • F02B75/04Engines with variable distances between pistons at top dead-centre positions and cylinder heads
    • F02B75/041Engines with variable distances between pistons at top dead-centre positions and cylinder heads by means of cylinder or cylinderhead positioning
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D13/00Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
    • F02D13/02Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D13/00Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
    • F02D13/02Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
    • F02D13/0223Variable control of the intake valves only
    • F02D13/0226Variable control of the intake valves only changing valve lift or valve lift and timing
    • F02D13/023Variable 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
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D13/00Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
    • F02D13/02Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
    • F02D13/0269Controlling the valves to perform a Miller-Atkinson cycle
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D15/00Varying compression ratio
    • F02D15/04Varying compression ratio by alteration of volume of compression space without changing piston stroke
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D41/00Electrical control of supply of combustible mixture or its constituents
    • F02D41/0002Controlling intake air
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L1/00Valve-gear or valve arrangements, e.g. lift-valve gear
    • F01L1/34Valve-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
    • F01L1/344Valve-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
    • F01L1/3442Valve-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
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L1/00Valve-gear or valve arrangements, e.g. lift-valve gear
    • F01L1/34Valve-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
    • F01L1/344Valve-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
    • F01L1/3442Valve-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
    • F01L2001/34423Details relating to the hydraulic feeding circuit
    • F01L2001/34426Oil control valves
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D13/00Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
    • F02D13/02Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
    • F02D13/0223Variable control of the intake valves only
    • F02D13/0234Variable control of the intake valves only changing the valve timing only
    • F02D13/0238Variable 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
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D41/00Electrical control of supply of combustible mixture or its constituents
    • F02D41/0002Controlling intake air
    • F02D2041/001Controlling intake air for engines with variable valve actuation
    • YGENERAL 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
    • Y02TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02TCLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00Road transport of goods or passengers
    • Y02T10/10Internal combustion engine [ICE] based vehicles
    • Y02T10/12Improving ICE efficiencies
    • YGENERAL 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
    • Y02TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02TCLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00Road transport of goods or passengers
    • Y02T10/10Internal combustion engine [ICE] based vehicles
    • Y02T10/40Engine management systems

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  • 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)

Description

本発明は、火花点火式内燃機関に関する。   The present invention relates to a spark ignition internal combustion engine.

機械圧縮比を変更可能な可変圧縮比機構と吸気弁の閉弁時期を制御可能な可変バルブタイミング機構とを具備し、機関中負荷運転時及び機関高負荷運転時には過給機による過給作用を行い、且つこれら機関中高負荷運転時においては実圧縮比を一定に保持した状態で機関負荷が低くなるにつれて機械圧縮比を増大すると共に吸気弁の閉弁時期を遅くするようにした火花点火式内燃機関が公知である(例えば特許文献1を参照)。   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. A spark ignition type internal combustion engine that increases the mechanical compression ratio and delays the closing timing of the intake valve as the engine load decreases while the actual compression ratio is kept constant during the medium and high load operation. The engine is known (see, for example, Patent Document 1).

特開2004−218522号公報JP 2004-218522 A

ところで、一般的に内燃機関では膨張比が大きくなればなるほど膨張行程時にピストンに対して押下げ力が作用する期間が長くなり、従って膨張比が大きくなるほど熱効率が向上する。従って、機関運転時における熱効率を向上させるためには機械圧縮比を可能な限り高くして膨張比を大きなものとすることが好ましい。   In general, in an internal combustion engine, the larger the expansion ratio, the longer the period during which the pressing force acts on the piston during the expansion stroke. Therefore, the greater the expansion ratio, the better the thermal efficiency. Therefore, in order to improve the thermal efficiency during engine operation, it is preferable to increase the mechanical compression ratio as much as possible to increase the expansion ratio.

ところが、このように膨張比を大きくすると、燃焼室内で発生した熱エネルギの多くが運動エネルギに変換されるため、排気ガスの温度が低くなる。また、これに伴って膨張行程末期における燃焼室内の排気ガスの圧力も低くなり、よって排気ガスは燃焼室から排出されにくくなる。このような傾向は、膨張比を20以上とした場合に特に顕著に現れる。   However, when the expansion ratio is increased in this way, most of the heat energy generated in the combustion chamber is converted into kinetic energy, so that the temperature of the exhaust gas is lowered. As a result, the pressure of the exhaust gas in the combustion chamber at the end of the expansion stroke is also lowered, so that the exhaust gas is hardly discharged from the combustion chamber. Such a tendency is particularly prominent when the expansion ratio is 20 or more.

一方、機関排気通路内に設けられる排気浄化触媒は一般的に或る温度以上にまで昇温されていないと良好な排気浄化作用を発揮することができない。このため、多くの内燃機関では機関本体から排出された排気ガスの熱により排気浄化触媒を高温に維持することとしている。   On the other hand, the exhaust purification catalyst provided in the engine exhaust passage generally cannot exhibit a good exhaust purification action unless the temperature is raised above a certain temperature. For this reason, in many internal combustion engines, the exhaust purification catalyst is maintained at a high temperature by the heat of the exhaust gas discharged from the engine body.

ところが、上述したように膨張比を大きくすると排気ガスの温度が低くなるため、単位流量当たりの排気ガスによって排気浄化触媒が昇温される温度が低くなる。また、膨張比を大きくすると排気ガスが燃焼室から排出されにくくなるため、排気浄化触媒に流入する排気ガスの流量が少なくなる。このため、膨張比が大きい状態で内燃機関を運転させると、排気浄化触媒の温度を高温に維持するのが困難になる。   However, as described above, when the expansion ratio is increased, the temperature of the exhaust gas is lowered, so that the temperature at which the exhaust purification catalyst is heated by the exhaust gas per unit flow rate is lowered. Further, if the expansion ratio is increased, the exhaust gas is not easily discharged from the combustion chamber, so the flow rate of the exhaust gas flowing into the exhaust purification catalyst is reduced. For this reason, when the internal combustion engine is operated with a large expansion ratio, it becomes difficult to maintain the temperature of the exhaust purification catalyst at a high temperature.

そこで、本発明の目的は、膨張比が大きい状態で内燃機関を運転させても、排気浄化触媒の温度を比較的高温に維持することができる火花点火式内燃機関を提供することにある。   Accordingly, an object of the present invention is to provide a spark ignition type internal combustion engine that can maintain the temperature of the exhaust purification catalyst at a relatively high temperature even when the internal combustion engine is operated in a state where the expansion ratio is large.

上記課題を解決するために、第1の発明では、機械圧縮比を変更可能な可変圧縮比機構と、実際の圧縮作用の開始時期を変更可能な実圧縮作用開始時期変更機構と、排気弁とを具備し、機関低負荷運転時には最大の膨張比が得られるように機械圧縮比を最大にし、上記最大の膨張比が20以上であり、更に機関低負荷運転時において排気弁の閉弁時期を吸気上死点の前後10°以内とした、火花点火式内燃機関が提供されるIn order to solve the above problems, in the first invention, a variable compression ratio mechanism capable of changing a mechanical compression ratio, an actual compression action start timing changing mechanism capable of changing an actual compression action start time, an exhaust valve, comprising a, to maximize the mechanical compression ratio so that the maximum expansion ratio is obtained at the time of engine low load operation, and at the maximum expansion ratio is 20 or more, a further closing timing of the exhaust valve at the time of engine low load operation Provided is a spark ignition internal combustion engine within 10 ° before and after intake top dead center .

上記課題を解決するために、第2の発明では、機械圧縮比を変更可能な可変圧縮比機構と、実際の圧縮作用の開始時期を変更可能な実圧縮作用開始時期変更機構と、排気弁の閉弁時期を変更可能な排気可変バルブタイミング機構とを具備し、機関低負荷運転時には最大の膨張比が得られるように機械圧縮比を最大にし、上記最大の膨張比が20以上であり、機関低負荷運転時において排気弁の閉弁時期の設定可能な領域が機関高負荷運転時よりも吸気上死点側に制限された、火花点火式内燃機関が提供されるIn order to solve the above problems, in the second invention, a variable compression ratio mechanism capable of changing a mechanical compression ratio, an actual compression action start timing changing mechanism capable of changing an actual compression action start timing, an exhaust valve ; and a changeable exhaust variable valve timing mechanism the closing timing, to maximize the mechanical compression ratio so that the maximum expansion ratio is obtained at the time of engine low load operation, and at the maximum expansion ratio is 20 or more, the engine Provided is a spark ignition type internal combustion engine in which a region in which the valve closing timing of an exhaust valve can be set during low load operation is limited to the intake top dead center side than during engine high load operation.

第3の発明では、第2の発明において、機関低負荷運転時において排気弁の閉弁時期を吸気上死点の前後10°以内とした。
第4の発明では、第2の発明において、吸気弁の開弁時期を変更可能な吸気可変バルブタイミング機構を更に具備し、機関低負荷運転時に吸気弁の開弁期間と排気弁の開弁期間とが重なるオーバーラップ期間が最小となるように排気弁の閉弁時期及び吸気弁の開弁時期が制御される。
第5の発明では、第2の発明において、吸気弁の開弁時期を変更可能な吸気可変バルブタイミング機構を更に具備し、機関低負荷運転時に吸気弁の開弁期間と排気弁の開弁期間とが重なるオーバーラップ期間がゼロとなるように排気弁の閉弁時期及び吸気弁の開弁時期が制御される。
In the third invention, in the second invention, the closing timing of the exhaust valve is set to be within 10 ° before and after the intake top dead center at the time of engine low load operation.
According to a fourth invention, in the second invention, an intake variable valve timing mechanism capable of changing the opening timing of the intake valve is further provided, and the intake valve opening period and the exhaust valve opening period during engine low load operation. The valve closing timing of the exhaust valve and the valve opening timing of the intake valve are controlled so that the overlap period in which the valve overlaps is minimized.
According to a fifth invention, in the second invention, an intake variable valve timing mechanism capable of changing the valve opening timing of the intake valve is further provided, and the intake valve opening period and the exhaust valve opening period during engine low load operation. The valve closing timing of the exhaust valve and the valve opening timing of the intake valve are controlled so that the overlap period in which they overlap is zero.

第6の発明では、第1又は第2の発明において、吸気弁の開弁時期を変更可能な吸気弁開弁時期変更機構を更に具備し、機関低負荷運転時には吸気弁の開弁時期を吸気上死点の前後10°以内とした。
第7の発明では、第1又は第2の発明において、機関低負荷運転時における実圧縮比が機関中高負荷運転時と同じ実圧縮比とされる。
第8の発明では、第7の発明において、機関低回転時には機関負荷に関わらずに上記実圧縮比が9〜11の範囲内とされる。
第9の発明では、第8の発明において、機関回転数が高くなるほど上記実圧縮比が高くされる。
In the sixth invention, the intake air in the first or second aspect, further capable of changing the intake valve opening timing changing mechanism the opening timing of the intake valve comprises a valve opening timing of the intake valve at the time of engine low load operation Within 10 ° before and after top dead center .
In the seventh invention, in the first or second aspect of the invention, the actual compression ratio at the time of engine low load operation is an engine medium and high load operation and the same actual compression ratio.
According to an eighth aspect, in the seventh aspect, the actual compression ratio is within a range of 9 to 11 regardless of the engine load at the time of low engine speed.
In a ninth aspect, in the eighth aspect, the actual compression ratio is increased as the engine speed increases.

第10の発明では、第1又は第2の発明において、上記実圧縮作用開始時期変更機構が吸気弁の閉弁時期を変更可能な吸気可変バルブタイミング機構からなる。
第11の発明では、第10の発明において、燃焼室内に供給される吸入空気量が吸気弁の閉弁時期を変えることによって制御される。
第12の発明では、第11の発明において、吸気弁の閉弁時期は、機関負荷が低くなるにつれて、燃焼室に供給される吸入空気量を制御し得る限界閉弁時期まで吸気下死点から離れる方向に移動せしめられる。
第13の発明では、第12の発明において、吸気弁の閉弁時期が上記限界閉弁時期に達したときの機関負荷よりも負荷の高い領域では燃焼室内に供給される吸入空気量が機関吸気通路内に配置されたスロットル弁によらずに吸気弁の閉弁時期を変えることによって制御される。
In a tenth aspect of the invention, in the first or second aspect of the invention, the actual compression action start timing changing mechanism comprises an intake variable valve timing mechanism capable of changing the closing timing of the intake valve.
In an eleventh aspect, in the tenth aspect, the amount of intake air supplied into the combustion chamber is controlled by changing the closing timing of the intake valve.
In the twelfth invention, in the eleventh invention, the closing timing of the intake valve is from the intake bottom dead center until the limit closing timing at which the amount of intake air supplied to the combustion chamber can be controlled as the engine load decreases. It can be moved away.
In a thirteenth aspect according to the twelfth aspect, in the twelfth aspect, the amount of intake air supplied into the combustion chamber is the engine intake air in a region where the load is higher than the engine load when the closing timing of the intake valve reaches the limit closing timing. It is controlled by changing the valve closing timing of the intake valve without depending on the throttle valve arranged in the passage.

第14の発明では、第13の発明において、吸気弁の閉弁時期が上記限界閉弁時期に達したときの機関負荷よりも負荷の高い領域ではスロットル弁が全開状態に保持される。
第15の発明では、第12の発明において、吸気弁の閉弁時期が上記限界閉弁時期に達したときの機関負荷よりも負荷の低い領域では機関吸気通路内に配置されたスロットル弁によって燃焼室内に供給される吸入空気量が制御される。
第16の発明では、第12の発明において、吸気弁の閉弁時期が上記限界閉弁時期に達したときの機関負荷よりも負荷の低い領域では負荷が低くなるほど空燃比が大きくされる。
In a fourteenth aspect, in the thirteenth aspect, the throttle valve is kept fully open in a region where the load is higher than the engine load when the closing timing of the intake valve reaches the limit closing timing.
According to a fifteenth aspect, in the twelfth aspect, combustion is performed by a throttle valve disposed in the engine intake passage in a region where the load is lower than the engine load when the valve closing timing of the intake valve reaches the limit valve closing timing. The amount of intake air supplied into the room is controlled.
In a sixteenth aspect, in the twelfth aspect, the air-fuel ratio is increased as the load decreases in a region where the load is lower than the engine load when the closing timing of the intake valve reaches the limit closing timing.

第17の発明では、第12の発明において、吸気弁の閉弁時期が上記限界閉弁時期に達したときの機関負荷よりも負荷の低い領域では吸気弁の閉弁時期が上記限界閉弁時期に保持される。
第18の発明では、第1又は第2の発明において、上記機械圧縮比は機関負荷が低くなるにつれて限界機械圧縮比まで増大せしめられる。
第19の発明では、第18の発明において、上記機械圧縮比が上記限界機械圧縮比に達したときの機関負荷よりも負荷の低い領域では機械圧縮比が上記限界機械圧縮比に保持される。
In a seventeenth aspect according to the twelfth aspect, in the twelfth aspect, in the region where the load is lower than the engine load when the closing timing of the intake valve reaches the limit closing timing, the closing timing of the intake valve is the limit closing timing. Retained.
In an eighteenth aspect based on the first or second aspect, the mechanical compression ratio is increased to a limit mechanical compression ratio as the engine load decreases.
In a nineteenth aspect, in the eighteenth aspect, the mechanical compression ratio is maintained at the limit mechanical compression ratio in a region where the load is lower than the engine load when the mechanical compression ratio reaches the limit mechanical compression ratio.

本発明によれば、できるだけ多くの排気ガスが燃焼室から排気浄化触媒へと排出されるため、膨張比が大きい状態で内燃機関を運転させても、排気浄化触媒の温度を比較的高温に維持することができる。   According to the present invention, as much exhaust gas as possible is discharged from the combustion chamber to the exhaust purification catalyst, so that the temperature of the exhaust purification catalyst is maintained at a relatively high temperature even when the internal combustion engine is operated with a large expansion ratio. can do.

図1に火花点火式内燃機関の側面断面図を示す。
図1を参照すると、1はクランクケース、2はシリンダブロック、3はシリンダヘッド、4はピストン、5は燃焼室、6は燃焼室5の頂面中央部に配置された点火プラグ、7は吸気弁、8は吸気ポート、9は排気弁、10は排気ポートをそれぞれ示す。吸気ポート8は吸気枝管11を介してサージタンク12に連結され、各吸気枝管11にはそれぞれ対応する吸気ポート8内に向けて燃料を噴射するための燃料噴射弁13が配置される。なお、燃料噴射弁13は各吸気枝管11に取付ける代りに各燃焼室5内に配置してもよい。
FIG. 1 shows a side sectional view of a spark ignition type internal combustion engine.
Referring to FIG. 1, 1 is a crankcase, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is a spark plug disposed at the center of the top surface of the combustion chamber 5, and 7 is an intake air 8 is an intake port, 9 is an exhaust valve, and 10 is an exhaust port. The intake port 8 is connected to a surge tank 12 via an intake branch pipe 11, and a fuel injection valve 13 for injecting fuel into the corresponding intake port 8 is arranged in each intake branch pipe 11. The fuel injection valve 13 may be arranged in each combustion chamber 5 instead of being attached to each intake branch pipe 11.

サージタンク12は吸気ダクト14を介して排気ターボチャージャ15のコンプレッサ15aの出口に連結され、コンプレッサ15aの入口は例えば熱線を用いた吸入空気量検出器16を介してエアクリーナ17に連結される。吸気ダクト14内にはアクチュエータ18によって駆動されるスロットル弁19が配置される。   The surge tank 12 is connected to the outlet of the compressor 15a of the exhaust turbocharger 15 via an intake duct 14, and the inlet of the compressor 15a is connected to an air cleaner 17 via an intake air amount detector 16 using, for example, heat rays. A throttle valve 19 driven by an actuator 18 is disposed in the intake duct 14.

一方、排気ポート10は排気マニホルド20を介して排気ターボチャージャ15の排気タービン15bの入口に連結され、排気タービン15bの出口は排気管21を介して排気浄化触媒を内蔵した触媒コンバータ22に連結される。排気管21内には空燃比センサ23が配置される。   On the other hand, the exhaust port 10 is connected to an inlet of an exhaust turbine 15b of an exhaust turbocharger 15 via an exhaust manifold 20, and an outlet of the exhaust turbine 15b is connected to a catalytic converter 22 containing an exhaust purification catalyst via an exhaust pipe 21. The An air-fuel ratio sensor 23 is disposed in the exhaust pipe 21.

一方、図1に示した実施形態では、クランクケース1とシリンダブロック2との連結部にクランクケース1とシリンダブロック2のシリンダ軸線方向の相対位置を変化させることによりピストン4が圧縮上死点に位置するときの燃焼室5の容積を変更可能な可変圧縮比機構Aが設けられており、また実際の圧縮作用の開始時期を変更するために吸気弁7の閉弁時期を制御可能であり且つ吸気弁7の開弁時期も個別に制御可能な吸気可変バルブタイミング機構Bが設けられており、更に排気弁7の開弁時期及び閉弁時期を個別に制御化能な排気可変バルブタイミング機構Cが設けられている。   On the other hand, in the embodiment shown in FIG. 1, the piston 4 is brought to the compression top dead center by changing the relative position of the crankcase 1 and the cylinder block 2 in the cylinder axis direction at the connecting portion between the crankcase 1 and the cylinder block 2. A variable compression ratio mechanism A capable of changing the volume of the combustion chamber 5 when positioned is provided, and the closing timing of the intake valve 7 can be controlled in order to change the actual start timing of the compression action, and An intake variable valve timing mechanism B capable of individually controlling the opening timing of the intake valve 7 is provided, and an exhaust variable valve timing mechanism C capable of individually controlling the opening timing and closing timing of the exhaust valve 7. Is provided.

電子制御ユニット30はデジタルコンピュータからなり、双方向性バス31によって互いに接続されたROM(リードオンリメモリ)32、RAM(ランダムアクセスメモリ)33、CPU(マイクロプロセッサ)34、入力ポート35および出力ポート36を具備する。吸入空気量検出器16の出力信号および空燃比センサ23の出力信号はそれぞれ対応するAD変換器37を介して入力ポート35に入力される。また、アクセルペダル40にはアクセルペダル40の踏込み量に比例した出力電圧を発生する負荷センサ41が接続され、負荷センサ41の出力電圧は対応するAD変換器37を介して入力ポート35に入力される。更に入力ポート35にはクランクシャフトが例えば30°回転する毎に出力パルスを発生するクランク角センサ42が接続される。一方、出力ポート36は対応する駆動回路38を介して点火プラグ6、燃料噴射弁13、スロットル弁駆動用アクチュエータ18、可変圧縮比機構Aおよび吸気可変バルブタイミング機構Bに接続される。   The electronic control unit 30 is composed of a digital computer, and is connected to each other by a bidirectional bus 31. A ROM (read only memory) 32, a RAM (random access memory) 33, a CPU (microprocessor) 34, an input port 35 and an output port 36. It comprises. The output signal of the intake air amount detector 16 and the output signal of the air-fuel ratio sensor 23 are input to the input port 35 via the corresponding AD converters 37, respectively. A load sensor 41 that generates an output voltage proportional to the amount of depression of the accelerator pedal 40 is connected to the accelerator pedal 40, and the output voltage of the load sensor 41 is input to the input port 35 via the corresponding AD converter 37. The Further, 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 18, the variable compression ratio mechanism A, and the intake variable valve timing mechanism B through corresponding drive circuits 38.

図2は図1に示す可変圧縮比機構Aの分解斜視図を示しており、図3は図解的に表した内燃機関の側面断面図を示している。図2を参照すると、シリンダブロック2の両側壁の下方には互いに間隔を隔てた複数個の突出部50が形成されており、各突出部50内にはそれぞれ断面円形のカム挿入孔51が形成されている。一方、クランクケース1の上壁面上には互いに間隔を隔ててそれぞれ対応する突出部50の間に嵌合せしめられる複数個の突出部52が形成されており、これらの各突出部52内にもそれぞれ断面円形のカム挿入孔53が形成されている。   2 is an exploded perspective view of the variable compression ratio mechanism A shown in FIG. 1, and FIG. 3 is a side sectional view of the internal combustion engine schematically shown. Referring to FIG. 2, a plurality of projecting portions 50 spaced from each other are formed below both side walls of the cylinder block 2, and cam insertion holes 51 each having a circular cross section are formed in each projecting portion 50. Has been. On the other hand, a plurality of protrusions 52 are formed on the upper wall surface of the crankcase 1 so as to be fitted between the corresponding protrusions 50 with a space between each other. Cam insertion holes 53 each having a circular cross section are formed.

図2に示したように一対のカムシャフト54,55が設けられており、各カムシャフト54,55上には一つおきに各カム挿入孔51内に回転可能に挿入される円形カム56が固定されている。これらの円形カム56は各カムシャフト54,55の回転軸線と共軸をなす。一方、各円形カム56間には図3においてハッチングで示すように各カムシャフト54,55の回転軸線に対して偏心配置された偏心軸57が延びており、この偏心軸57上に別の円形カム58が偏心して回転可能に取付けられている。図2に示したようにこれら円形カム58は各円形カム56間に配置されており、これら円形カム58は対応する各カム挿入孔53内に回転可能に挿入されている。   As shown in FIG. 2, a pair of camshafts 54 and 55 are provided, and on each camshaft 54 and 55, a circular cam 56 is rotatably inserted into each cam insertion hole 51. It is fixed. These circular cams 56 are coaxial with the rotational axes of the camshafts 54 and 55. On the other hand, an eccentric shaft 57 arranged eccentrically with respect to the rotation axis of each camshaft 54, 55 extends between the circular cams 56 as shown by hatching in FIG. A cam 58 is eccentrically mounted for rotation. As shown in FIG. 2, the circular cams 58 are disposed between the circular cams 56, and the circular cams 58 are rotatably inserted into the corresponding cam insertion holes 53.

図3(A)に示すような状態から各カムシャフト54,55上に固定された円形カム56を図3(A)において実線の矢印で示したように互いに反対方向に回転させると偏心軸57が下方中央に向けて移動するために円形カム58がカム挿入孔53内において図3(A)の破線の矢印に示すように円形カム56とは反対方向に回転し、図3(B)に示したように偏心軸57が下方中央まで移動すると円形カム58の中心が偏心軸57の下方へ移動する。   When the circular cams 56 fixed on the camshafts 54 and 55 are rotated in opposite directions as indicated by solid arrows in FIG. 3A from the state shown in FIG. 3 moves toward the lower center, so that the circular cam 58 rotates in the opposite direction to the circular cam 56 in the cam insertion hole 53 as shown by the broken arrow in FIG. As shown, when the eccentric shaft 57 moves to the lower center, the center of the circular cam 58 moves below the eccentric shaft 57.

図3(A)と図3(B)とを比較するとわかるようにクランクケース1とシリンダブロック2の相対位置は円形カム56の中心と円形カム58の中心との距離によって定まり、円形カム56の中心と円形カム58の中心との距離が大きくなるほどシリンダブロック2はクランクケース1から離れる。シリンダブロック2がクランクケース1から離れるとピストン4が圧縮上死点に位置するときの燃焼室5の容積は増大し、従って各カムシャフト54,55を回転させることによってピストン4が圧縮上死点に位置するときの燃焼室5の容積を変更することができる。   3A and 3B, the relative positions of the crankcase 1 and the cylinder block 2 are determined by the distance between the center of the circular cam 56 and the center of the circular cam 58. The cylinder block 2 moves away from the crankcase 1 as the distance between the center and the center of the circular cam 58 increases. When the cylinder block 2 moves away from the crankcase 1, the volume of the combustion chamber 5 increases when the piston 4 is positioned at the compression top dead center. Therefore, by rotating the camshafts 54 and 55, the piston 4 is compressed at the top dead center. The volume of the combustion chamber 5 when it is located at can be changed.

図2に示したように各カムシャフト54,55をそれぞれ反対方向に回転させるために駆動モータ59の回転軸にはそれぞれ螺旋方向が逆向きの一対のウォームギア61,62が取付けられており、これらウォームギア61,62と噛合する歯車63,64がそれぞれ各カムシャフト54,55の端部に固定されている。この実施形態では駆動モータ59を駆動することによってピストン4が圧縮上死点に位置するときの燃焼室5の容積を広い範囲に亘って変更することができる。なお、図1から図3に示した可変圧縮比機構Aは一例を示すものであっていかなる形式の可変圧縮比機構でも用いることができる。   As shown in FIG. 2, in order to rotate the camshafts 54 and 55 in opposite directions, a pair of worm gears 61 and 62 having opposite spiral directions are attached to the rotation shaft of the drive motor 59, respectively. Gears 63 and 64 that mesh with the worm gears 61 and 62 are fixed to end portions of the camshafts 54 and 55, respectively. In this embodiment, by driving the drive motor 59, the volume of the combustion chamber 5 when the piston 4 is located at the compression top dead center can be changed over a wide range. The variable compression ratio mechanism A shown in FIGS. 1 to 3 shows an example, and any type of variable compression ratio mechanism can be used.

一方、図4は図1において吸気弁7を駆動するためのカムシャフト70に対して設けられている吸気可変バルブタイミング機構Bを示している。図4に示したように吸気可変バルブタイミング機構Bはカムシャフト70の一端に取付けられてカムシャフト70のカムの位相を変更するためのカム位相変更部B1と、カムシャフト70と吸気弁7のバルブリフタ24との間に配置されてカムシャフト70のカムの作用角を異なる作用角に変更して吸気弁7に伝達するカム作用角変更部B2から構成されている。なお、カム作用角変更部B2については図4に側面断面図と平面図とが示されている。   On the other hand, FIG. 4 shows an intake variable valve timing mechanism B provided for the camshaft 70 for driving the intake valve 7 in FIG. As shown in FIG. 4, the intake variable valve timing mechanism B is attached to one end of the camshaft 70 to change the cam phase of the camshaft 70, the cam phase changing portion B1, and the camshaft 70 and the intake valve 7. The cam operating angle changing portion B2 is disposed between the valve lifter 24 and changes the cam operating angle of the camshaft 70 to a different operating angle and transmits the cam operating angle to the intake valve 7. As for the cam working angle changing portion B2, a side sectional view and a plan view are shown in FIG.

まず初めに吸気可変バルブタイミング機構Bのカム位相変更部B1について説明すると、このカム位相変更部B1は機関のクランク軸によりタイミングベルトを介して矢印方向に回転せしめられるタイミングプーリ71と、タイミングプーリ71と一緒に回転する円筒状ハウジング72と、カムシャフト70と一緒に回転し且つ円筒状ハウジング72に対して相対回転可能な回転軸73と、円筒状ハウジング72の内周面から回転軸73の外周面まで延びる複数の仕切壁74と、各仕切壁74の間で回転軸73の外周面から円筒状ハウジング72の内周面まで延びるベーン75とを具備しており、各ベーン75の両側にはそれぞれ進角用油圧室76と遅角用油圧室77とが形成されている。   First, the cam phase changing unit B1 of the intake variable valve timing mechanism B will be described. The cam phase changing unit B1 is rotated by the crankshaft of the engine in the direction of the arrow through the timing belt, and the timing pulley 71. A cylindrical housing 72 that rotates together with the camshaft 70, a rotating shaft 73 that rotates together with the camshaft 70 and that can rotate relative to the cylindrical housing 72, and an outer periphery of the rotating shaft 73 from the inner peripheral surface of the cylindrical housing 72. A plurality of partition walls 74 extending to the surface, and vanes 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 are provided on both sides of each vane 75. An advance hydraulic chamber 76 and a retard hydraulic chamber 77 are formed respectively.

各油圧室76,77への作動油の供給制御は作動油供給制御弁78によって行われる。この作動油供給制御弁78は各油圧室76,77にそれぞれ連結された油圧ポート79,80と、油圧ポンプ81から吐出された作動油の供給ポート82と、一対のドレインポート83,84と、各ポート79,80,82,83,84間の連通遮断制御を行うスプール弁85とを具備している。   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.

カムシャフト70のカムの位相を進角すべきときは図4においてスプール弁85が下方に移動せしめられ、供給ポート82から供給された作動油が油圧ポート79を介して進角用油圧室76に供給されると共に遅角用油圧室77内の作動油がドレインポート84から排出される。このとき回転軸73は円筒状ハウジング72に対して矢印X方向に相対回転せしめられる。   When the cam phase of the camshaft 70 is to be advanced, the spool valve 85 is moved downward in FIG. 4, and the hydraulic oil supplied from the supply port 82 enters the advance angle hydraulic chamber 76 via the hydraulic port 79. The hydraulic oil in the retarding hydraulic chamber 77 is supplied and discharged from the drain port 84. At this time, the rotation shaft 73 is rotated relative to the cylindrical housing 72 in the arrow X direction.

これに対し、カムシャフト70のカムの位相を遅角すべきときは図4においてスプール弁85が上方に移動せしめられ、供給ポート82から供給された作動油が油圧ポート80を介して遅角用油圧室77に供給されると共に進角用油圧室76内の作動油がドレインポート83から排出される。このとき回転軸73は円筒状ハウジング72に対して矢印Xと反対方向に相対回転せしめられる。   On the other hand, when the cam phase of the camshaft 70 is to be retarded, the spool valve 85 is moved upward in FIG. 4 and the hydraulic oil supplied from the supply port 82 is used for retarding via the hydraulic port 80. The hydraulic oil in the advance hydraulic chamber 76 is discharged from the drain port 83 while being supplied to the hydraulic chamber 77. At this time, the rotating shaft 73 is rotated relative to the cylindrical housing 72 in the direction opposite to the arrow X.

回転軸73が円筒状ハウジング72に対して相対回転せしめられているときにスプール弁85が図4に示した中立位置に戻されると回転軸73の相対回転動作は停止せしめられ、回転軸73はそのときの相対回転位置に保持される。従ってカム位相変更部B1によってカムシャフト70のカムの位相を所望の量だけ進角又は遅角させることができる。即ち、カム位相変更部B1によって吸気弁7の開弁時期を任意に進角又は遅角させることができることになる。   If the spool valve 85 is returned to the neutral position shown in FIG. 4 while the rotation shaft 73 is rotated relative to the cylindrical housing 72, the relative rotation operation of the rotation shaft 73 is stopped, and the rotation shaft 73 The relative rotation position at that time is held. Therefore, the cam phase changing portion B1 can advance or retard the cam phase of the camshaft 70 by a desired amount. That is, the valve opening timing of the intake valve 7 can be arbitrarily advanced or retarded by the cam phase changing unit B1.

次に吸気可変バルブタイミング機構Bのカム作用角変更部B2について説明すると、このカム作用角変更部B2はカムシャフト70と平行に並列配置され且つアクチュエータ91によって軸線方向に移動せしめられる制御ロッド90と、カムシャフト70のカム92と係合し且つ制御ロッド90上に形成された軸線方向に延びるスプライン93に摺動可能に嵌合せしめられている中間カム94と、吸気弁7を駆動するためにバルブリフタ24と係合し且つ制御ロッド90上に形成された螺旋状に延びるスプライン95に摺動可能に嵌合する揺動カム96とを具備しており、揺動カム96上にはカム97が形成されている。   Next, the cam working angle changing portion B2 of the intake variable valve timing mechanism B will be described. The cam working angle changing portion B2 is arranged in parallel with the camshaft 70 and is moved in the axial direction by the actuator 91. An intermediate cam 94 slidably engaged with an axially extending spline 93 formed on the control rod 90 and engaged with the cam 92 of the camshaft 70, and for driving the intake valve 7 A swing cam 96 that engages with the valve lifter 24 and slidably engages with a spiral extending spline 95 formed on the control rod 90 is provided. Is formed.

カムシャフト70が回転するとカム92によって中間カム94が常に一定の角度だけ揺動せしめられ、このとき揺動カム96も一定の角度だけ揺動せしめられる。一方、中間カム94及び揺動カム96は制御ロッド90の軸線方向には移動不能に支持されており、従って制御ロッド90がアクチュエータ91によって軸線方向に移動せしめられたときに揺動カム96は中間カム94に対して相対回転せしめられることになる。   When the camshaft 70 is rotated, the intermediate cam 94 is always swung by a certain angle by the cam 92. At this time, the rocking cam 96 is also swung by a certain angle. On the other hand, the intermediate cam 94 and the swing cam 96 are supported so as not to move in the axial direction of the control rod 90. Therefore, when the control rod 90 is moved in the axial direction by the actuator 91, the swing cam 96 is intermediate. It is rotated relative to the cam 94.

中間カム94と揺動カム96との相対回転位置関係によりカムシャフト70のカム92が中間カム94と係合し始めたときに揺動カム96のカム97がバルブリフタ24と係合し始める場合には図5(B)においてaで示したように吸気弁7の開弁期間及びリフトは最も大きくなる。これに対し、アクチュエータ91によって揺動カム96が中間カム94に対して図4の矢印Y方向に相対回転せしめられると、カムシャフト70のカム92が中間カム94に係合した後、暫らくしてから揺動カム96のカム97がバルブリフタ24と係合する。この場合には図5(B)においてbで示したように吸気弁7の開弁期間及びリフト量はaに比べて小さくなる。   When the cam 97 of the swing cam 96 starts to engage with the valve lifter 24 when the cam 92 of the cam shaft 70 starts to engage with the intermediate cam 94 due to the relative rotational positional relationship between the intermediate cam 94 and the swing cam 96. As shown by a in FIG. 5B, the valve opening period and lift of the intake valve 7 become the largest. On the other hand, when the swing cam 96 is rotated relative to the intermediate cam 94 in the direction of the arrow Y by the actuator 91, the cam 92 of the camshaft 70 is engaged with the intermediate cam 94 for a while. Thereafter, the cam 97 of the swing cam 96 is engaged with the valve lifter 24. In this case, as indicated by b in FIG. 5B, the valve opening period and the lift amount of the intake valve 7 are smaller than a.

揺動カム96が中間カム94に対して図4の矢印Y方向に更に相対回転せしめられると図5(B)においてcで示したように吸気弁7の開弁期間及びリフト量は更に小さくなる。即ち、アクチュエータ91により中間カム94と揺動カム96の相対回転位置を変更することによって吸気弁7の開弁期間を任意に変えることができる。ただし、この場合、吸気弁7のリフト量は吸気弁7の開弁期間が短くなるほど小さくなる。   When the swing cam 96 is further rotated relative to the intermediate cam 94 in the direction of the arrow Y in FIG. 4, the valve opening period and the lift amount of the intake valve 7 are further reduced as indicated by c in FIG. . That is, the valve opening period of the intake valve 7 can be arbitrarily changed by changing the relative rotational position of the intermediate cam 94 and the swing cam 96 by the actuator 91. However, in this case, the lift amount of the intake valve 7 becomes smaller as the opening period of the intake valve 7 becomes shorter.

このようにカム位相変更部B1によって吸気弁7の開弁時期を任意に変更することができ、カム作用角変更部B2によって吸気弁7の開弁期間を任意に変更することができるのでカム位相変更部B1とカム作用角変更部B2との双方によって、即ち吸気可変バルブタイミング機構Bによって吸気弁7の開弁時期と開弁期間とを、即ち吸気弁7の開弁時期と閉弁時期とを任意に変更することができることになる。   Thus, the cam phase changing unit B1 can arbitrarily change the valve opening timing of the intake valve 7, and the cam operating angle changing unit B2 can arbitrarily change the valve opening period of the intake valve 7, so that the cam phase By both the change part B1 and the cam working angle change part B2, that is, by the intake variable valve timing mechanism B, the valve opening timing and the valve opening period of the intake valve 7, that is, the valve opening timing and the valve closing timing of the intake valve 7 Can be arbitrarily changed.

なお、図1および図4に示した吸気可変バルブタイミング機構Bは一例を示すものであって、図1および図4に示した例以外の種々の形式の可変バルブタイミング機構を用いることができる。   The intake variable valve timing mechanism B shown in FIGS. 1 and 4 shows an example, and various types of variable valve timing mechanisms other than the examples shown in FIGS. 1 and 4 can be used.

また、排気可変バルブタイミング機構Cも、基本的に吸気可変バルブタイミング機構Bと同様な構成を有し、排気弁9の開弁時期と開弁期間とを、即ち排気弁9の開弁時期と閉弁時期とを任意に変更することができる。   The exhaust variable valve timing mechanism C has basically the same configuration as the intake variable valve timing mechanism B, and the opening timing and opening period of the exhaust valve 9, that is, the opening timing of the exhaust valve 9 are The valve closing timing can be arbitrarily changed.

次に図6を参照しつつ本願において使用されている用語の意味について説明する。なお、図6の(A),(B),(C)には説明のために燃焼室容積が50mlでピストンの行程容積が500mlであるエンジンが示されており、これら図6の(A),(B),(C)において燃焼室容積とはピストンが圧縮上死点に位置するときの燃焼室の容積を表している。   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.

図6(A)は機械圧縮比について説明している。機械圧縮比は圧縮行程時のピストンの行程容積と燃焼室容積のみから機械的に定まる値であってこの機械圧縮比は(燃焼室容積+行程容積)/燃焼室容積で表される。図6(A)に示した例ではこの機械圧縮比は(50ml+500ml)/50ml=11となる。   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.

図6(B)は実圧縮比について説明している。この実圧縮比は実際に圧縮作用が開始されたときからピストンが上死点に達するまでの実際のピストン行程容積と燃焼室容積から定まる値であってこの実圧縮比は(燃焼室容積+実際の行程容積)/燃焼室容積で表される。即ち、図6(B)に示したように圧縮行程においてピストンが上昇を開始しても吸気弁が開弁している間は圧縮作用は行われず、吸気弁が閉弁したときから実際の圧縮作用が開始される。従って実圧縮比は実際の行程容積を用いて上記のように表される。図6(B)に示した例では実圧縮比は(50ml+450ml)/50ml=10となる。   FIG. 6B describes the actual compression ratio. This actual compression ratio is a value determined from the actual piston stroke volume and the combustion chamber volume from when the compression action is actually started until the piston reaches top dead center, and this actual compression ratio is (combustion chamber volume + actual (Stroke volume) / combustion chamber volume. That is, as shown in FIG. 6B, even if the piston starts to rise in the compression stroke, the compression action is not performed while the intake valve is open, and the actual compression 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.

図6(C)は膨張比について説明している。膨張比は膨張行程時のピストンの行程容積と燃焼室容積から定まる値であってこの膨張比は(燃焼室容積+行程容積)/燃焼室容積で表される。図6(C)に示した例ではこの膨張比は(50ml+500ml)/50ml=11となる。   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.

次に図7および図8を参照しつつ本発明において最も基本となっている特徴について説明する。なお、図7は理論熱効率と膨張比との関係を示しており、図8は本発明において負荷に応じ使い分けられている通常のサイクルと超高膨張比サイクルとの比較を示している。   Next, the most basic features of 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.

図8(A)は吸気弁が下死点近傍で閉弁し、ほぼ圧縮下死点付近からピストンによる圧縮作用が開始される場合の通常のサイクルを示している。この図8(A)に示す例でも図6の(A),(B),(C)に示す例と同様に燃焼室容積が50mlとされ、ピストンの行程容積が500mlとされている。図8(A)からわかるように通常のサイクルでは機械圧縮比は(50ml+500ml)/50ml=11であり、実圧縮比もほぼ11であり、膨張比も(50ml+500ml)/50ml=11となる。即ち、通常の内燃機関では機械圧縮比と実圧縮比と膨張比とがほぼ等しくなる。   FIG. 8A shows a normal cycle when the intake valve closes near the bottom dead center and the compression action by the piston is started from the vicinity of the compression bottom dead center. In the example shown in FIG. 8A, the combustion chamber volume is 50 ml, and the stroke volume of the piston is 500 ml, as in the examples shown in FIGS. 6A, 6B, and 6C. As can be seen from FIG. 8A, in a normal cycle, the mechanical compression ratio is (50 ml + 500 ml) / 50 ml = 11, the actual compression ratio is almost 11, and the expansion ratio is also (50 ml + 500 ml) / 50 ml = 11. That is, in a normal internal combustion engine, the mechanical compression ratio, the actual compression ratio, and the expansion ratio are substantially equal.

図7における実線は実圧縮比と膨張比とがほぼ等しい場合の、即ち通常のサイクルにおける理論熱効率の変化を示している。この場合には膨張比が大きくなるほど、即ち実圧縮比が高くなるほど理論熱効率が高くなることがわかる。従って通常のサイクルにおいて理論熱効率を高めるには実圧縮比を高くすればよいことになる。しかしながら機関高負荷運転時におけるノッキングの発生の制約により実圧縮比は最大でも12程度までしか高くすることができず、斯くして通常のサイクルにおいては理論熱効率を十分に高くすることはできない。   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.

一方、このような状況下で本発明者は機械圧縮比と実圧縮比とを厳密に区分して理論熱効率を高めることについて検討し、その結果理論熱効率は膨張比が支配し、理論熱効率に対して実圧縮比はほとんど影響を与えないことを見出したのである。即ち、実圧縮比を高くすると爆発力は高まるが圧縮するために大きなエネルギが必要となり、斯くして実圧縮比を高めても理論熱効率はほとんど高くならない。   On the other hand, in this situation, the present inventor has studied to increase the theoretical thermal efficiency by strictly dividing the mechanical compression ratio and the actual compression ratio, and as a result, the theoretical thermal efficiency is governed by the expansion ratio, and the theoretical thermal efficiency It was found that the actual compression ratio has 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.

これに対し、膨張比を大きくすると膨張行程時にピストンに対し押下げ力が作用する期間が長くなり、斯くしてピストンがクランクシャフトに回転力を与えている期間が長くなる。従って膨張比は大きくすれば大きくするほど理論熱効率が高くなる。図7の破線は実圧縮比を10に固定した状態で膨張比を高くしていった場合の理論熱効率を示している。このように実圧縮比を低い値に維持した状態で膨張比を高くしたときの理論熱効率の上昇量と、図7の実線で示すように実圧縮比も膨張比と共に増大せしめられる場合の理論熱効率の上昇量とは大きな差がないことがわかる。   On the other hand, when the expansion ratio is increased, the period during which the pressing force acts on the piston during the expansion stroke becomes longer, and thus the period during which the piston applies the rotational force to the crankshaft becomes longer. Therefore, the larger the expansion ratio, the higher the theoretical thermal efficiency. The broken line in FIG. 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 maintaining the actual compression ratio at a low value, and the theoretical thermal efficiency when the actual compression ratio is increased with the expansion ratio as shown by the solid line in FIG. It can be seen that there is no significant difference from the amount of increase.

このように実圧縮比が低い値に維持されているとノッキングが発生することがなく、従って実圧縮比を低い値に維持した状態で膨張比を高くするとノッキングの発生を阻止しつつ理論熱効率を大巾に高めることができる。図8(B)は可変圧縮比機構A及び吸気可変バルブタイミング機構Bを用いて、実圧縮比を低い値に維持しつつ膨張比を高めるようにした場合の一例を示している。   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 the intake variable valve timing mechanism B are used to increase the expansion ratio while maintaining the actual compression ratio at a low value.

図8(B)を参照すると、この例では可変圧縮比機構Aにより燃焼室容積が50mlから20mlまで減少せしめられる。一方、吸気可変バルブタイミング機構Bによって実際のピストン行程容積が500mlから200mlになるまで吸気弁の閉弁時期が遅らされる。その結果、この例では実圧縮比は(20ml+200ml)/20ml=11となり、膨張比は(20ml+500ml)/20ml=26となる。図8(A)に示した通常のサイクルでは上述したように実圧縮比がほぼ11で膨張比が11であり、この場合に比べると図8(B)に示した場合には膨張比のみが26まで高められていることがわかる。これが超高膨張比サイクルと称される所以である。   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 intake valve closing timing is delayed by the intake variable valve timing mechanism B until the actual piston stroke volume is reduced from 500 ml to 200 ml. As a result, in this example, the actual compression ratio is (20 ml + 200 ml) / 20 ml = 11, and the expansion ratio is (20 ml + 500 ml) / 20 ml = 26. In the normal cycle shown in FIG. 8A, the actual compression ratio is almost 11 and the expansion ratio is 11 as described above. Compared to this case, only the expansion ratio is shown in FIG. 8B. It can be seen that it has been increased to 26. This is why it is called an ultra-high expansion ratio cycle.

上述したように一般的に言って内燃機関では機関負荷が低いほど熱効率が悪くなり、従って車両走行時における熱効率を向上させるためには、即ち燃費を向上させるには機関低負荷運転時における熱効率を向上させることが必要となる。一方、図8(B)に示した超高膨張比サイクルでは圧縮行程時の実際のピストン行程容積が小さくされるために燃焼室5内に吸入し得る吸入空気量は少なくなり、従ってこの超高膨張比サイクルは機関負荷が比較的低いときにしか採用できないことになる。従って本発明では機関低負荷運転時には図8(B)に示す超高膨張比サイクルとし、機関高負荷運転時には図8(A)に示す通常のサイクルとするようにしている。これが本発明の基本としている特徴である。   As described above, in general, in an internal combustion engine, the lower the engine load, the worse the thermal efficiency. Therefore, in order to improve the thermal efficiency when the vehicle is running, that is, to improve the fuel efficiency, the thermal efficiency during the engine low load operation is reduced. It is necessary to improve. On the other hand, in the ultra-high expansion ratio cycle shown in FIG. 8B, since the actual piston stroke volume during the compression stroke is reduced, the amount of intake air that can be sucked into the combustion chamber 5 is reduced. The expansion ratio cycle can only be adopted when the engine load is relatively low. Therefore, in the present invention, the super high expansion ratio cycle shown in FIG. 8 (B) is used during low engine load operation, and the normal cycle shown in FIG. 8 (A) is used during high engine load operation. This is the basic feature of the present invention.

図9は機関回転数の低い定常運転時における運転制御全般について示している。以下この図9を参照しつつ運転制御全般について説明する。
図9には機関負荷に応じた機械圧縮比、膨張比、吸気弁7の閉弁時期、実圧縮比、吸入空気量、スロットル弁17の開度およびポンピング損失の各変化が示されている。なお、本実施形態では触媒コンバータ22内の三元触媒によって排気ガス中の未燃HC,COおよびNOxを同時に低減し得るように、通常、燃焼室5内における平均空燃比は空燃比センサ23の出力信号に基づいて理論空燃比にフィードバック制御されている。
FIG. 9 shows the overall operation control during steady operation at a low engine speed. The overall operation control will be described below with reference to FIG.
FIG. 9 shows changes in the mechanical compression ratio, the expansion ratio, the closing timing of the intake valve 7, the actual compression ratio, the intake air amount, the opening degree of the throttle valve 17 and the pumping loss according to the engine load. In the present embodiment, the average air-fuel ratio in the combustion chamber 5 is usually the air-fuel ratio sensor 23 so that unburned HC, CO and NO x in the exhaust gas can be simultaneously reduced by the three-way catalyst in the catalytic converter 22. Feedback control to the stoichiometric air-fuel ratio based on the output signal.

さて、上述したように機関高負荷運転時には図8(A)に示した通常のサイクルが実行される。従って図9に示したように、このときには機械圧縮比が低くされるために膨張比は低く、図9において実線で示したように吸気弁7の閉弁時期は早められている。また、このときには吸入空気量は多く、このときスロットル弁17の開度は全開又はほぼ全開に保持されているのでポンピング損失は零となっている。   As described above, the normal cycle shown in FIG. 8A is executed during engine high load operation. Therefore, as shown in FIG. 9, at this time, the mechanical compression ratio is lowered, so the expansion ratio is low, and the closing timing of the intake valve 7 is advanced as shown by the solid line in FIG. 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 or almost fully open, so that the pumping loss is zero.

一方、図9に示したように機関負荷が低くなるとそれに伴って機械圧縮比が増大され、従って膨張比も増大される。またこのときには実圧縮比がほぼ一定に保持されるように図9において実線で示したように機関負荷が低くなるにつれて吸気弁7の閉弁時期が遅くされる。なお、このときにもスロットル弁17は全開又はほぼ全開状態に保持されており、従って燃焼室5内に供給される吸入空気量はスロットル弁17によらずに吸気弁7の閉弁時期を変えることによって制御されている。このときにもポンピング損失は零となる。   On the other hand, as shown in FIG. 9, when the engine load is lowered, the mechanical compression ratio is increased accordingly, and thus the expansion ratio is also increased. Further, at this time, the valve closing timing of the intake valve 7 is delayed as the engine load becomes lower as shown by the solid line in FIG. 9 so that the actual compression ratio is kept substantially constant. At this time as well, the throttle valve 17 is kept fully open or substantially fully open, and therefore the amount of intake air supplied into the combustion chamber 5 changes the closing timing of the intake valve 7 regardless of the throttle valve 17. Is controlled by that. Also at this time, the pumping loss becomes zero.

このように機関高負荷運転状態から機関負荷が低くなるときには実圧縮比がほぼ一定のもとで吸入空気量が減少するにつれて機械圧縮比が増大せしめられる。即ち、吸入空気量の減少に比例してピストン4が圧縮上死点に達したときの燃焼室5の容積が減少せしめられる。従ってピストン4が圧縮上死点に達したときの燃焼室5の容積は吸入空気量に比例して変化していることになる。なお、このとき燃焼室5内の空燃比は理論空燃比となっているのでピストン4が圧縮上死点に達したときの燃焼室5の容積は燃料量に比例して変化していることになる。   As described above, when the engine load is reduced from the engine high load operation state, the mechanical compression ratio is increased as the intake air amount is decreased while the actual compression ratio is substantially constant. That is, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center is decreased in proportion to the decrease in the intake air amount. Therefore, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the intake air amount. At this time, since the air-fuel ratio in the combustion chamber 5 is the stoichiometric air-fuel ratio, the volume of the combustion chamber 5 when the piston 4 reaches the compression top dead center changes in proportion to the fuel amount. Become.

機関負荷が更に低くなると機械圧縮比は更に増大せしめられ、機械圧縮比が燃焼室5の構造上限界となる限界機械圧縮比に達すると、機械圧縮比が限界機械圧縮比に達したときの機関負荷L1よりも負荷の低い領域では機械圧縮比が限界機械圧縮比に保持される。従って機関低負荷運転時には機械圧縮比は最大となり、膨張比も最大となる。別の言い方をすると本発明では機関低負荷運転時に最大の膨張比が得られるように機械圧縮比が最大にされる。また、このとき実圧縮比は機関中高負荷運転時とほぼ同じ実圧縮比に維持される。 When the engine load is further reduced, the mechanical compression ratio is further increased. When the mechanical compression ratio reaches the limit mechanical compression ratio that is the structural limit of the combustion chamber 5, the engine when the mechanical compression ratio reaches the limit mechanical compression ratio is reached. In the region where the load is lower than the load L 1 , the mechanical compression ratio is maintained at the limit mechanical compression ratio. Therefore, the mechanical compression ratio is maximized and the expansion ratio is maximized during engine low load operation. In other words, in the present invention, the mechanical compression ratio is maximized so that the maximum expansion ratio can be obtained during engine low load operation. Further, at this time, the actual compression ratio is maintained at substantially the same actual compression ratio as that during the engine medium and high load operation.

一方、図9において実線で示したように吸気弁7の閉弁時期は機関負荷が低くなるにつれて燃焼室5内に供給される吸入空気量を制御し得る限界閉弁時期まで遅らされ、吸気弁7の閉弁時期が限界閉弁時期に達したときの機関負荷L2よりも負荷の低い領域では吸気弁7の閉弁時期が限界閉弁時期に保持される。吸気弁7の閉弁時期が限界閉弁時期に保持されるともはや吸気弁7の閉弁時期の変化によっては吸入空気量を制御し得ないので他の何らかの方法によって吸入空気量を制御する必要がある。 On the other hand, as shown by the solid line in FIG. 9, the closing timing of the intake valve 7 is delayed to the limit closing timing that can control the amount of intake air supplied into the combustion chamber 5 as the engine load becomes lower. closing timing of the valve 7 is the closing timing of the intake valve 7 in the region of a load lower than the engine load L 2 when reaching the limit closing timing is held at the limit closing timing. If the closing timing of the intake valve 7 is held at the limit closing timing, the intake air amount can no longer be controlled by the change in the closing timing of the intake valve 7, so it is necessary to control the intake air amount by some other method There is.

図9に示した実施形態ではこのとき、即ち吸気弁7の閉弁時期が限界閉弁時期に達したときの機関負荷L2よりも負荷の低い領域ではスロットル弁17によって燃焼室5内に供給される吸入空気量が制御される。ただし、スロットル弁17による吸入空気量の制御が行われると図9に示したようにポンピング損失が増大する。 In the embodiment shown in FIG. 9, at this time, that is, in a region where the load is lower than the engine load L 2 when the closing timing of the intake valve 7 has reached the limit closing timing, the throttle valve 17 supplies the fuel into the combustion chamber 5. The amount of intake air to be controlled is controlled. However, if the intake air amount is controlled by the throttle valve 17, the pumping loss increases as shown in FIG.

なお、このようなポンピング損失が発生しないように吸気弁7の閉弁時期が限界閉弁時期に達したときの機関負荷L2よりも負荷の低い領域ではスロットル弁17を全開又はほぼ全開に保持した状態で機関負荷が低くなるほど空燃比を大きくすることもできる。このときには燃料噴射弁13を燃焼室5内に配置して成層燃焼させることが好ましい。 In order to prevent such a pumping loss, the throttle valve 17 is kept fully open or almost fully open in a region where the load is lower than the engine load L 2 when the closing timing of the intake valve 7 reaches the limit closing timing. In this state, the air-fuel ratio can be increased as the engine load decreases. At this time, it is preferable to arrange the fuel injection valve 13 in the combustion chamber 5 and perform stratified combustion.

図9に示したように機関低回転時には機関負荷にかかわらずに実圧縮比がほぼ一定に保持される。このときの実圧縮比は機関中高負荷運転時の実圧縮比に対してほぼ±10パーセントの範囲内とされ、好ましくは±5パーセントの範囲内とされる。なお、本実施形態では機関低回転時の実圧縮比はほぼ10±1、即ち、9から11の間とされる。ただし、機関回転数が高くなると燃焼室5内の混合気に乱れが発生するためにノッキングが発生しづらくなり、従って本発明による実施形態では機関回転数が高くなるほど実圧縮比が高くされる。   As shown in FIG. 9, the actual compression ratio is maintained substantially constant regardless of the engine load at the time of low engine speed. The actual compression ratio at this time is approximately within a range of ± 10%, preferably within a range of ± 5% with respect to the actual compression ratio at the time of engine medium and high load operation. In this embodiment, the actual compression ratio at the time of low engine rotation is approximately 10 ± 1, that is, between 9 and 11. However, when the engine speed increases, the air-fuel mixture in the combustion chamber 5 is disturbed, so that knocking does not easily occur. Therefore, in the embodiment according to the present invention, the actual compression ratio increases as the engine speed increases.

一方、上述したように図8(B)に示す超高膨張比サイクルでは膨張比が26とされる。この膨張比は高いほど好ましいが20以上であればかなり高い理論熱効率を得ることができる。従って本発明では膨張比が20以上となるように可変圧縮比機構Aが形成されている。   On the other hand, as described above, the expansion ratio is 26 in the ultra-high expansion ratio cycle shown in FIG. This expansion ratio is preferably as high as possible, but if it is 20 or more, a considerably high theoretical thermal efficiency can be obtained. Therefore, in the present invention, the variable compression ratio mechanism A is formed so that the expansion ratio is 20 or more.

また、図9に示した例では機械圧縮比は機関負荷に応じて連続的に変化せしめられている。しかしながら機械圧縮比は機関負荷に応じて段階的に変化させることもできる。   In the example shown in FIG. 9, the mechanical compression ratio is continuously changed according to the engine load. However, the mechanical compression ratio can be changed stepwise according to the engine load.

一方、図9において破線で示すように機関負荷が低くなるにつれて吸気弁7の閉弁時期を早めることによってもスロットル弁17によらずに吸入空気量を制御することができる。従って、図9において実線で示した場合と破線で示した場合とをいずれも包含し得るように表現すると、本実施形態では吸気弁7の閉弁時期は、機関負荷が低くなるにつれて、燃焼室内に供給される吸入空気量を制御し得る限界閉弁時期L2まで吸気下死点BDCから離れる方向に移動せしめられることになる。 On the other hand, as shown by the broken line in FIG. 9, the intake air amount can be controlled without depending on the throttle valve 17 by advancing the closing timing of the intake valve 7 as the engine load becomes lower. Therefore, if expressed in FIG. 9 so as to include both the case indicated by the solid line and the case indicated by the broken line, in this embodiment, the valve closing timing of the intake valve 7 is set in the combustion chamber as the engine load decreases. limit closing timing capable of controlling the amount of intake air fed to L 2 will be moved in the direction away from intake bottom dead center BDC to.

次に、図8(B)に示した超高膨張比サイクルが実行される低負荷運転時に焦点を当てて排気弁9の閉弁時期について説明する。
一般に、超高膨張比サイクルが実行される低負荷運転時は燃焼室5内での混合気の燃焼による熱の発生量自体が少ないことから、燃焼室5から排出される排気ガスの温度が低くなり易い。これに加えて、内燃機関では膨張比が大きくなればなるほど膨張行程時にピストンに対して押下げ力が作用する期間が長くなるため、燃焼室内での混合気の燃焼によって生じた熱エネルギの多くがピストンの運動エネルギに変換され、これに伴って膨張行程末期における燃焼室5内の燃焼ガスの温度は低くなる。このため、図8(B)に示した超高膨張比サイクルが実行されているときには排気行程時に燃焼室5から排気マニホルド20へと排出される排気ガスの温度は極めて低いものとなる。このような傾向は、膨張比を20以上とした場合に特に顕著に現れ、膨張比を20以上とした超高膨張比サイクルの実行時と膨張比を12程度とした通常のサイクルとでは燃焼室5から排出される排気ガスの温度が約100℃程度異なる。
Next, the closing timing of the exhaust valve 9 will be described focusing on the low load operation in which the super high expansion ratio cycle shown in FIG. 8B is executed.
In general, during a low load operation in which an ultra-high expansion ratio cycle is performed, the amount of heat generated by combustion of the air-fuel mixture in the combustion chamber 5 is small, so the temperature of the exhaust gas discharged from the combustion chamber 5 is low. Easy to be. In addition, in the internal combustion engine, the larger the expansion ratio, the longer the period during which the pressing force acts on the piston during the expansion stroke, so that much of the heat energy generated by the combustion of the air-fuel mixture in the combustion chamber is reduced. This is converted into kinetic energy of the piston, and accordingly, the temperature of the combustion gas in the combustion chamber 5 at the end of the expansion stroke is lowered. For this reason, when the ultra-high expansion ratio cycle shown in FIG. 8B is executed, the temperature of the exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 during the exhaust stroke is extremely low. Such a tendency is particularly noticeable when the expansion ratio is set to 20 or more. The combustion chamber is used when an ultra-high expansion ratio cycle with an expansion ratio of 20 or more is executed and when a normal cycle with an expansion ratio of about 12 is used. The temperature of the exhaust gas discharged from 5 differs by about 100 ° C.

一方、多くの内燃機関では排気ガス中に含まれる有害成分(例えば、HC、CO、NOX等)を浄化するため、機関排気通路内に酸化触媒、三元触媒、NOX吸蔵還元触媒等の排気浄化触媒が設けられている。このような排気浄化触媒は、その温度が活性温度以上にならないと排気ガス中の有害成分を効果的に浄化することができない。ここで、多くの内燃機関では排気ガスの温度が活性温度よりもかなり高いため、排気ガスを排気浄化触媒に流入させることによって排気浄化触媒の温度を活性温度以上に維持するようにしている。 On the other hand, in many internal combustion engines, in order to purify harmful components (for example, HC, CO, NO x, etc.) contained in the exhaust gas, an oxidation catalyst, a three-way catalyst, a NO x storage reduction catalyst, etc. are disposed in the engine exhaust passage. An exhaust purification catalyst is provided. Such an exhaust purification catalyst cannot effectively purify harmful components in the exhaust gas unless the temperature becomes higher than the activation temperature. Here, in many internal combustion engines, since the temperature of the exhaust gas is considerably higher than the activation temperature, the temperature of the exhaust purification catalyst is maintained at or above the activation temperature by flowing the exhaust gas into the exhaust purification catalyst.

ところが、図8(B)に示した超高膨張比サイクルが実行されると、燃焼室5から排気マニホルド20へと排出される排気ガスの温度は活性温度よりも僅かにしか高くならないことから、排気ガスを排気浄化触媒に流入させても排気浄化触媒の温度をその活性温度以上に維持することが困難となる。従って、超高膨張比サイクルが実行されているときに排気浄化触媒の温度をその活性温度以上に維持するためには、できるだけ多くの排気ガスを排気浄化触媒に流入させる必要がある。   However, when the ultra-high expansion ratio cycle shown in FIG. 8B is executed, the temperature of the exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 is only slightly higher than the activation temperature. Even if the exhaust gas flows into the exhaust purification catalyst, it becomes difficult to maintain the temperature of the exhaust purification catalyst at or above its activation temperature. Therefore, in order to maintain the temperature of the exhaust purification catalyst at the activation temperature or higher when the ultra-high expansion ratio cycle is being executed, it is necessary to flow as much exhaust gas as possible into the exhaust purification catalyst.

ここで、図10を参照して排気弁9の閉弁時期と燃焼室5から排気マニホルド20へと排出される排気ガスの流量との関係について考えてみる。図10(A)は排気弁9をほぼ吸気上死点において閉弁した場合、図10(B)は排気弁9を吸気上死点よりも早く閉弁した場合、図10(C)は排気弁9を吸気上死点よりも遅く閉弁した場合の排気弁9及び吸気弁7のリフト変化をそれぞれ示している。   Now, with reference to FIG. 10, consider the relationship between the closing timing of the exhaust valve 9 and the flow rate of the exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20. FIG. 10A shows a case where the exhaust valve 9 is closed substantially at the intake top dead center, FIG. 10B shows a case where the exhaust valve 9 is closed earlier than the intake top dead center, and FIG. The lift changes of the exhaust valve 9 and the intake valve 7 when the valve 9 is closed later than the intake top dead center are shown.

図10(B)に示したように、排気弁9を吸気上死点よりも早く閉弁した場合、排気弁9の閉弁時における燃焼室5の容積は、ピストンが吸気上死点に位置する場合の燃焼室の容積(燃焼室容積)よりも大きく、排気弁9の閉弁後には閉弁時の燃焼室5の容積に相当する排気ガスが燃焼室5内に残ることになる。このため、排気弁9の閉弁後にも比較的多くの排気ガスが燃焼室5内に残ることになり、よって燃焼室5内の排気ガスを十分に排気マニホルド20へと排出することができず、排気浄化触媒に流入する排気ガスの流量が少ないものとなる。   As shown in FIG. 10B, when the exhaust valve 9 is closed earlier than the intake top dead center, the volume of the combustion chamber 5 when the exhaust valve 9 is closed is that the piston is positioned at the intake top dead center. When the exhaust valve 9 is closed, the exhaust gas corresponding to the volume of the combustion chamber 5 when the valve is closed remains in the combustion chamber 5 after the exhaust valve 9 is closed. For this reason, a relatively large amount of exhaust gas remains in the combustion chamber 5 even after the exhaust valve 9 is closed, so that the exhaust gas in the combustion chamber 5 cannot be sufficiently discharged to the exhaust manifold 20. Therefore, the flow rate of the exhaust gas flowing into the exhaust purification catalyst is small.

一方、図10(C)に示したように、排気弁9を吸気上死点よりも遅く閉弁した場合、吸気上死点においても排気弁9が開かれていることから、ピストン4が吸気上死点に到達したときに燃焼室5内の排気ガスのほとんどは排気ポート10内に流出する。ところが、吸気上死点以降においても排気弁9が開かれていると、ピストン4の下降に伴って、一旦排気ポート10内へ流出した排気ガスの一部が再び燃焼室5内に流入してしまう。   On the other hand, as shown in FIG. 10C, when the exhaust valve 9 is closed later than the intake top dead center, the exhaust valve 9 is opened even at the intake top dead center. When the top dead center is reached, most of the exhaust gas in the combustion chamber 5 flows out into the exhaust port 10. However, if the exhaust valve 9 is opened even after the intake top dead center, a part of the exhaust gas once flowing into the exhaust port 10 flows into the combustion chamber 5 again as the piston 4 descends. End up.

特に、超高膨張比サイクルが実行されているときには、膨張行程時において燃焼室5内の燃焼ガスがかなり膨張することから、膨張行程末期における燃焼ガスの圧力は比較的低い。このため、排気行程において燃焼室5から排気ポート10に流出する排気ガスの勢いは弱く、よって吸気上死点到達後にピストン4が下降すると一旦排気ポート10へ流出した排気ガスの一部が再び燃焼室5内に流入し易くなる。   In particular, when the ultra-high expansion ratio cycle is being executed, the combustion gas in the combustion chamber 5 expands considerably during the expansion stroke, so the pressure of the combustion gas at the end of the expansion stroke is relatively low. For this reason, the momentum of the exhaust gas flowing out from the combustion chamber 5 to the exhaust port 10 in the exhaust stroke is weak. Therefore, when the piston 4 descends after reaching the intake top dead center, a part of the exhaust gas that has flowed out to the exhaust port 10 once again burns It becomes easy to flow into the chamber 5.

このように、排気弁9を吸気上死点よりも遅く閉弁した場合には、一旦排気ポート10に流出した排気ガスが再び燃焼室5内に戻ってくるため、燃焼室内5の排気ガスを十分に排気マニホルド20へと排出することができず、排気浄化触媒に流入する排気ガスの流量が少ないものとなる。   As described above, when the exhaust valve 9 is closed later than the intake top dead center, the exhaust gas that has once flowed out to the exhaust port 10 returns to the combustion chamber 5 again. The exhaust gas cannot be sufficiently discharged to the exhaust manifold 20, and the flow rate of the exhaust gas flowing into the exhaust purification catalyst becomes small.

そこで、本実施形態では、図8(B)に示した超高膨張比サイクルが実行されるとき、すなわち機械圧縮比が高いときには、排気弁9の閉弁時期が吸気上死点よりも早過ぎたり遅過ぎたりすることのないように、排気弁9の閉弁時期の設定可能な領域を吸気上死点側に制限することとしている。   Therefore, in this embodiment, when the ultra-high expansion ratio cycle shown in FIG. 8B is executed, that is, when the mechanical compression ratio is high, the closing timing of the exhaust valve 9 is too early than the intake top dead center. In order to prevent the exhaust valve 9 from being too late, the region in which the valve closing timing of the exhaust valve 9 can be set is limited to the intake top dead center side.

図11は、機械圧縮比に応じた排気弁9の閉弁時期の設定可能な領域を示す図である。図11に示したように、排気弁9の設定可能な領域は設定可能な最大進角量と最大遅角量との間の領域となっている。図から分かるように、排気弁9の閉弁時期の設定可能な最大進角量は機械圧縮比が高くなるほど小さく(遅く)され、逆に排気弁9の閉弁時期の設定可能な最大遅角量は機械圧縮比が高くなるほど小さく(早く)される。このため、排気弁9の閉弁時期の設定可能な領域は、機械圧縮比が高くなるほど小さく、すなわち機械圧縮比が高くなるほど制限されている。例えば、図11に示したように、機械圧縮比が低いときには排気弁9の閉弁時期の設定可能な領域はΔTOC1であるのに対して機械圧縮比が高いときには排気弁9の閉弁時期の設定可能な領域はΔTOC2(ΔTOC2<ΔTOC1)とされる。   FIG. 11 is a diagram showing a region in which the valve closing timing of the exhaust valve 9 can be set according to the mechanical compression ratio. As shown in FIG. 11, the settable region of the exhaust valve 9 is a region between the settable maximum advance amount and maximum retard amount. As can be seen from the figure, the maximum advance angle that can be set for the closing timing of the exhaust valve 9 is smaller (slower) as the mechanical compression ratio becomes higher, and conversely, the maximum delay that can be set for the closing timing of the exhaust valve 9 The amount is reduced (faster) as the mechanical compression ratio increases. For this reason, the region in which the valve closing timing of the exhaust valve 9 can be set is smaller as the mechanical compression ratio is higher, that is, the region is limited as the mechanical compression ratio is higher. For example, as shown in FIG. 11, when the mechanical compression ratio is low, the region where the closing timing of the exhaust valve 9 can be set is ΔTOC1, whereas when the mechanical compression ratio is high, the closing timing of the exhaust valve 9 is The settable area is ΔTOC2 (ΔTOC2 <ΔTOC1).

或いは、図8(B)に示した超高膨張比サイクルが実行されるとき、すなわち機械圧縮比が高いときには、排気弁9の閉弁時期が吸気上死点よりも早過ぎたり遅過ぎたりすることを確実に防止すべく、図10(A)に示したように排気弁9の閉弁時期をほぼ吸気上死点にするようにしてもよい。   Alternatively, when the ultra-high expansion ratio cycle shown in FIG. 8B is executed, that is, when the mechanical compression ratio is high, the closing timing of the exhaust valve 9 may be earlier or later than the intake top dead center. In order to surely prevent this, as shown in FIG. 10A, the valve closing timing of the exhaust valve 9 may be set substantially at the intake top dead center.

このように、機械圧縮比が高いときには排気弁9の閉弁時期の設定可能な領域を吸気上死点側に制限したり、或いは排気弁9の閉弁時期をほぼ吸気上死点にしたりすることにより、燃焼室5内の排気ガスを十分に排気マニホルド20へと排出して、排気浄化触媒に流入する排気ガスの流量を多いものとすることができる。   As described above, when the mechanical compression ratio is high, the region where the closing timing of the exhaust valve 9 can be set is limited to the intake top dead center side, or the closing timing of the exhaust valve 9 is almost set to the intake top dead center. As a result, the exhaust gas in the combustion chamber 5 can be sufficiently discharged to the exhaust manifold 20, and the flow rate of the exhaust gas flowing into the exhaust purification catalyst can be increased.

すなわち、排気弁9は吸気上死点付近で閉弁せしめられるため、図10(B)に示したように排気弁9を吸気上死点よりも早く閉弁した場合に比べて、排気弁9の閉弁時における燃焼室5の容積が小さく、よって排気弁9の閉弁後に燃焼室5内に残る排気ガスの量を少なくすることができる。また、排気弁9が吸気上死点付近で閉弁せしめられるため、図10(C)に示したように排気弁9を吸気上死点よりも遅く閉弁した場合に比べて、排気ポート10内へ流出した排気ガスのうち燃焼室5内に再び流入する排気ガスの量を少なくすることができる。このため、図10(A)に示したように排気弁9が吸気上死点付近で閉弁させた場合には、図10(B)及び(C)に示したように排気弁9を吸気上死点から離れて閉弁させた場合に比べて、燃焼室5内の排気ガスを十分に排気マニホルド20へと排出して、排気浄化触媒に流入する排気ガスの流量を多いものとすることができる。その結果、超高膨張比サイクルが実行される低負荷運転時においても排気浄化触媒を活性温度以上に維持することができるようになる。   That is, since the exhaust valve 9 is closed near the intake top dead center, as shown in FIG. 10B, the exhaust valve 9 is closed compared to the case where the exhaust valve 9 is closed earlier than the intake top dead center. The volume of the combustion chamber 5 when the valve is closed is small, so that the amount of exhaust gas remaining in the combustion chamber 5 after the exhaust valve 9 is closed can be reduced. Further, since the exhaust valve 9 is closed in the vicinity of the intake top dead center, the exhaust port 10 is compared with the case where the exhaust valve 9 is closed later than the intake top dead center as shown in FIG. The amount of exhaust gas that flows back into the combustion chamber 5 out of the exhaust gas that flows into the interior can be reduced. Therefore, when the exhaust valve 9 is closed in the vicinity of the intake top dead center as shown in FIG. 10A, the exhaust valve 9 is inhaled as shown in FIGS. 10B and 10C. Compared to the case where the valve is closed away from the top dead center, the exhaust gas in the combustion chamber 5 is sufficiently discharged to the exhaust manifold 20 and the flow rate of the exhaust gas flowing into the exhaust purification catalyst is increased. Can do. As a result, the exhaust purification catalyst can be maintained at the activation temperature or higher even during low load operation in which the ultra-high expansion ratio cycle is executed.

なお、上記「ほぼ吸気上死点」との記載は、吸気上死点の前後約10°以内、好ましくは吸気上死点の前後約5°以内を示している。   Note that the description of “substantially intake top dead center” indicates within about 10 ° before and after the intake top dead center, preferably within about 5 ° before and after the intake top dead center.

また、機械圧縮比を高くすると吸気上死点における燃焼室容積が小さくなり、斯くして排気弁9の閉弁時期によっては排気弁9がピストン4と干渉してしまう。   Further, when the mechanical compression ratio is increased, the combustion chamber volume at the intake top dead center is reduced, and thus the exhaust valve 9 interferes with the piston 4 depending on the closing timing of the exhaust valve 9.

図10には排気弁9又は吸気弁7がピストン4と干渉する限界を示すピストン干渉ラインが示されており、排気弁9のリフト曲線がピストン干渉ラインと交錯すると排気弁9はピストン4と干渉することになる。ここで、図10(C)では、排気弁9のリフト曲線がピストン干渉ラインと交錯している。このことは、排気弁9を吸気上死点よりも遅く閉弁した場合、遅くした程度にもよるが、排気弁9とピストン4とが干渉してしまうことを意味している。   FIG. 10 shows a piston interference line indicating the limit at which the exhaust valve 9 or the intake valve 7 interferes with the piston 4. When the lift curve of the exhaust valve 9 intersects with the piston interference line, the exhaust valve 9 interferes with the piston 4. Will do. Here, in FIG. 10C, the lift curve of the exhaust valve 9 intersects with the piston interference line. This means that when the exhaust valve 9 is closed later than the intake top dead center, the exhaust valve 9 and the piston 4 interfere with each other, depending on the degree of delay.

これに対して、本実施形態によれば、機械圧縮比が高いときには排気弁9の閉弁時期の設定可能な領域が吸気上死点側に制限され、特に排気弁9の閉弁時期の設定可能な最大遅角量が小さくされる。このため、図10(A)に示したように、機械圧縮比が高くなっても排気弁9がピストン4と干渉するのが防止される。   On the other hand, according to this embodiment, when the mechanical compression ratio is high, the region where the closing timing of the exhaust valve 9 can be set is limited to the intake top dead center side, and in particular, the closing timing of the exhaust valve 9 is set. The maximum possible retard amount is reduced. For this reason, as shown in FIG. 10A, the exhaust valve 9 is prevented from interfering with the piston 4 even when the mechanical compression ratio is increased.

ところで、吸気弁7の開弁期間と排気弁9の開弁期間とが重なるバルブオーバーラップがある場合、その期間に応じても燃焼室5内から排気マニホルド20へと排出される排気ガスの量が変化する。以下、図12を参照して吸気弁7の開弁期間と排気弁9の開弁期間とが重なるオーバーラップ期間と燃焼室5から排気マニホルド20へと排出される排気ガスの量との関係について考える。図12(A)はオーバーラップ期間がゼロである場合、図12(B)はオーバーラップ期間が大きい場合の排気弁9及び吸気弁7のリフト変化をそれぞれ示している。   By the way, when there is a valve overlap in which the valve opening period of the intake valve 7 and the valve opening period of the exhaust valve 9 overlap, the amount of exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 in accordance with the valve overlap period. Changes. Hereinafter, with reference to FIG. 12, the relationship between the overlap period in which the valve opening period of the intake valve 7 and the valve opening period of the exhaust valve 9 overlap and the amount of exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 will be described. Think. FIG. 12A shows the lift changes of the exhaust valve 9 and the intake valve 7 when the overlap period is zero, and FIG. 12B shows the lift changes of the exhaust valve 9 and the intake valve 7 when the overlap period is long.

一般に吸気弁7と排気弁9とが同時に開いていると、燃焼室5内の排気ガスの一部や一旦燃焼室5から排気ポート10へ流出した排気ガスの一部が吸気ポート8内に流入することがある。このように、排気ガスの一部が吸気ポート8内に流入すると、その分だけ燃焼室5から排気マニホルド20へと排出される排気ガスが少なくなる。   In general, when the intake valve 7 and the exhaust valve 9 are open at the same time, a part of the exhaust gas in the combustion chamber 5 and a part of the exhaust gas once flowing out from the combustion chamber 5 to the exhaust port 10 flow into the intake port 8. There are things to do. Thus, when a part of the exhaust gas flows into the intake port 8, the amount of exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 is reduced accordingly.

従って、図12(B)に示したようにオーバーラップ期間が大きい場合には、排気ガスが吸気ポート8内へ多量に流入することが多く、よって燃焼室5から排気マニホルド20へと排出される排気ガスが少なくなることが多い。このため、このような場合には、排気浄化触媒に流入する排気ガスの流量が少なくなる。   Accordingly, when the overlap period is long as shown in FIG. 12B, a large amount of exhaust gas often flows into the intake port 8 and is thus discharged from the combustion chamber 5 to the exhaust manifold 20. Exhaust gas often decreases. For this reason, in such a case, the flow rate of the exhaust gas flowing into the exhaust purification catalyst is reduced.

そこで、本実施形態では、図8(B)に示した超高膨張比サイクルが実行されるとき、すなわち機械圧縮比が高いときには、図12(A)に示したようにオーバーラップ期間が設定可能な範囲のうち最小になるように排気弁7の閉弁時期及び吸気弁9の開弁時期を制御することとしている。従って、例えば、設定可能なオーバーラップ期間が10°〜60°となっている内燃機関では機械圧縮比が高いときにはオーバーラップ期間が10°とされ、設定可能なオーバーラップ期間が0°〜50°となっている内燃機関では機械圧縮比が高いときにはオーバーラップ期間は0°とされる。   Therefore, in this embodiment, when the ultra-high expansion ratio cycle shown in FIG. 8B is executed, that is, when the mechanical compression ratio is high, the overlap period can be set as shown in FIG. The valve closing timing of the exhaust valve 7 and the valve opening timing of the intake valve 9 are controlled so as to be the smallest in this range. Therefore, for example, in an internal combustion engine having a settable overlap period of 10 ° to 60 °, the overlap period is set to 10 ° when the mechanical compression ratio is high, and the settable overlap period is set to 0 ° to 50 °. In the internal combustion engine, the overlap period is set to 0 ° when the mechanical compression ratio is high.

このように、機械圧縮比が高いときにオーバーラップ期間を最小とすることにより、吸気ポート8内へ流入する排気ガスが少なくなるため、燃焼室5から排気マニホルド20へと排出される排気ガスは多くなり、よって排気浄化触媒に流入する排気ガスの流量が多くなる。   Thus, by minimizing the overlap period when the mechanical compression ratio is high, the exhaust gas flowing into the intake port 8 is reduced, so that the exhaust gas discharged from the combustion chamber 5 to the exhaust manifold 20 is Therefore, the flow rate of exhaust gas flowing into the exhaust purification catalyst increases.

なお、機械圧縮比が高いときのオーバーラップ期間は、機械圧縮比が低いときのオーバーラップ期間よりも短ければ必ずしも最小でなくてもよい。従って、例えば、機械圧縮比が高いときのオーバーラップ期間は、設定可能な範囲のうち最小でなくても10°以下であればよい。   The overlap period when the mechanical compression ratio is high may not necessarily be the minimum as 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 may be 10 ° or less even if it is not the minimum in the settable range.

また、上述したように機械圧縮比を高くすると吸気上死点における燃焼室容積が小さくなり、斯くして吸気弁7の開弁時期によっては吸気弁7がピストン4と干渉してしまう。   Further, as described above, when the mechanical compression ratio is increased, the combustion chamber volume at the intake top dead center is reduced. Therefore, the intake valve 7 interferes with the piston 4 depending on the opening timing of the intake valve 7.

図12には排気弁9又は吸気弁7がピストン4と干渉する限界を示すピストン干渉ラインが示されており、吸気弁7のリフト曲線がピストン干渉ラインと交錯すると吸気弁7はピストン4と干渉することになる。ここで、図12(B)では、吸気弁7のリフト曲線がピストン干渉ラインと交錯している。このことは、オーバーラップ期間を大きくすると、吸気弁7とピストン4とが干渉してしまうことを意味している。すなわち、本実施形態では、上述したように排気弁9の閉弁時期がほぼ吸気上死点とされるため、オーバーラップ期間が大きいことは吸気弁7の開弁時期が大きく進角せしめられることを意味する。このように吸気弁7の開弁時期が大きく進角せしめられると、吸気弁7とピストン4とが干渉してしまうことになる。   FIG. 12 shows a piston interference line indicating the limit at which the exhaust valve 9 or the intake valve 7 interferes with the piston 4. When the lift curve of the intake valve 7 intersects with the piston interference line, the intake valve 7 interferes with the piston 4. Will do. Here, in FIG. 12B, the lift curve of the intake valve 7 intersects with the piston interference line. This means that if the overlap period is increased, the intake valve 7 and the piston 4 interfere with each other. That is, in this embodiment, as described above, the closing timing of the exhaust valve 9 is almost the intake top dead center, so that the large overlap period means that the opening timing of the intake valve 7 is greatly advanced. Means. Thus, if the valve opening timing of the intake valve 7 is greatly advanced, the intake valve 7 and the piston 4 interfere with each other.

これに対して、本実施形態によれば、機械圧縮比が高いときにはオーバーラップ期間が最小とされるため、吸気弁7の開弁時期がほぼ吸気上死点又はそれ以降とされる。このため、図12(A)に示したように、機械圧縮比が高くなっても吸気弁7がピストンと干渉するのが防止される。   On the other hand, according to the present embodiment, when the mechanical compression ratio is high, the overlap period is minimized, so that the valve opening timing of the intake valve 7 is substantially at or near the intake top dead center. For this reason, as shown in FIG. 12A, the intake valve 7 is prevented from interfering with the piston even when the mechanical compression ratio becomes high.

図13に、本実施形態の火花点火式内燃機関の運転制御の制御ルーチンを示す。図13を参照するとまず初めにステップ101において機関負荷L及び機関回転数Neが取得される。次いで、ステップ102において図14(A)に示すマップから目標実圧縮比が算出される。図14(A)に示したようにこの目標実圧縮比は機関回転数Neが高くなるほど高くなる。次いで、ステップ103では図14(B)に示したマップから機械圧縮比CRが算出される。即ち、実圧縮比を目標実圧縮比とするのに必要な機械圧縮比CRが機関負荷L及び機関回転数Neの関数として図14(B)に示したようなマップの形で予めROM32内に記憶されており、このマップから機械圧縮比CRが算出される。   FIG. 13 shows a control routine for operation control of the spark ignition type internal combustion engine of the present embodiment. Referring to FIG. 13, first, at step 101, the engine load L and the engine speed Ne are acquired. Next, in step 102, the target actual compression ratio is calculated from the map shown in FIG. As shown in FIG. 14A, the target actual compression ratio increases as the engine speed Ne increases. Next, at step 103, the mechanical compression ratio CR is calculated from the map shown in FIG. That is, the mechanical compression ratio CR required to make the actual compression ratio the target actual compression ratio is stored in advance in the ROM 32 as a function of the engine load L and the engine speed Ne in the form of a map as shown in FIG. The mechanical compression ratio CR is calculated from this map.

更に、要求吸入空気量を燃焼室5内に供給するのに必要な吸気弁7の閉弁時期ICが機関負荷L及び機関回転数Neの関数として図14(C)に示したようなマップの形で予めROM32内に記憶されている。ステップ104では、このマップから吸気弁7の閉弁時期ICが算出される。   Further, the valve closing timing IC of the intake valve 7 necessary for supplying the required intake air amount into the combustion chamber 5 is a map as shown in FIG. 14C as a function of the engine load L and the engine speed Ne. Is stored in the ROM 32 in advance. In step 104, the valve closing timing IC of the intake valve 7 is calculated from this map.

次いで、ステップ105では、機関負荷Lが所定値L3よりも小さいか否かが判定される。ここで、所定値L3は、例えば、それ以上機関負荷が小さくなると排気ガスの温度の低下に伴って排気浄化触媒の温度がその活性温度以下に低下してしまう可能性が有る機関負荷に等しい値とされる。ステップ105において、機関負荷Lが所定値L3よりも小さいと判定された場合にはステップ106へと進む。ステップ106では排気弁9の閉弁時期ECがほぼ吸気上死点とされ、次いでステップ107ではオーバーラップ期間ΔOLが最小とされ、ステップ110へと進む。 Then, in step 105, the engine load L is whether less than a predetermined value L 3 is determined. Here, for example, the predetermined value L 3 is equal to an engine load in which if the engine load is further reduced, the temperature of the exhaust purification catalyst may be lowered below its activation temperature as the temperature of the exhaust gas decreases. Value. If it is determined in step 105 that the engine load L is smaller than the predetermined value L 3 , the process proceeds to step 106. In step 106, the valve closing timing EC of the exhaust valve 9 is set to the intake top dead center, and then in step 107, the overlap period ΔOL is minimized and the routine proceeds to step 110.

一方、ステップ105において、機関負荷が所定値L3以上であると判定された場合にはステップ108へと進む。ステップ108では図15(A)に示したマップから排気弁9の閉弁時期ECが算出され、次いでステップ109では図15(B)に示したマップからオーバーラップ期間ΔOLが算出される。即ち、排気弁9の閉弁時期EC及びオーバーラップ期間ΔOLが機関負荷L及び機関回転数Neの関数として図15(A)、(B)に示したようなマップの形で予めROM32内に記憶されており、このマップから排気弁9の閉弁時期EC及びオーバーラップ期間ΔOLが算出される。その後、ステップ110へと進む。 On the other hand, if it is determined in step 105 that the engine load is equal to or greater than the predetermined value L 3 , the process proceeds to step 108. In step 108, the valve closing timing EC of the exhaust valve 9 is calculated from the map shown in FIG. 15A. Next, in step 109, the overlap period ΔOL is calculated from the map shown in FIG. That is, the valve closing timing EC and the overlap period ΔOL of the exhaust valve 9 are stored in advance in the ROM 32 as a function of the engine load L and the engine speed Ne in the form of a map as shown in FIGS. From this map, the valve closing timing EC and the overlap period ΔOL of the exhaust valve 9 are calculated. Thereafter, the process proceeds to step 110.

ステップ110では、機械圧縮比が機械圧縮比CRとなるように可変圧縮比機構Aが制御され、吸気弁7の閉弁時期が閉弁時期ICとなるように且つオーバーラップ期間がオーバーラップ期間ΔOLとなるように吸気可変バルブタイミング機構Bが制御される。更に、排気弁9の閉弁時期が閉弁時期ECとなるように排気可変バルブタイミング機構Cが制御される。   In step 110, the variable compression ratio mechanism A is controlled so that the mechanical compression ratio becomes the mechanical compression ratio CR, the closing timing of the intake valve 7 becomes the closing timing IC, and the overlapping period is the overlapping period ΔOL. The intake variable valve timing mechanism B is controlled so that Further, the exhaust variable valve timing mechanism C is controlled so that the closing timing of the exhaust valve 9 becomes the closing timing EC.

火花点火式内燃機関の全体図である。1 is an overall view of a spark ignition internal combustion engine. 可変圧縮比機構の分解斜視図である。It is a disassembled perspective view of a variable compression ratio mechanism. 図解的に表した内燃機関の側面断面図である。1 is a schematic side sectional view of an internal combustion engine. 可変バルブタイミング機構を示す図である。It is a figure which shows a variable valve timing mechanism. 吸気弁および排気弁のリフト量を示す図である。It is a figure which shows the lift amount of an intake valve and an exhaust valve. 機械圧縮比、実圧縮比および膨張比を説明するための図である。It is a figure for demonstrating a mechanical compression ratio, an actual compression ratio, and an expansion ratio. 理論熱効率と膨張比との関係を示す図である。It is a figure which shows the relationship between theoretical thermal efficiency and an expansion ratio. 通常のサイクルおよび超高膨張比サイクルを説明するための図である。It is a figure for demonstrating a normal cycle and a super-high expansion ratio cycle. 機関負荷に応じた機械圧縮比等の変化を示す図である。It is a figure which shows changes, such as a mechanical compression ratio according to an engine load. 吸気弁及び排気弁のリフト変化を示す図である。It is a figure which shows the lift change of an intake valve and an exhaust valve. 機械圧縮比に応じた排気弁の閉弁時期の設定可能な領域を示す図である。It is a figure which shows the area | region which can set the valve closing timing of an exhaust valve according to mechanical compression ratio. 吸気弁及び排気弁のリフト変化を示す図である。It is a figure which shows the lift change of an intake valve and an exhaust valve. 運転制御を行うためのフローチャートである。It is a flowchart for performing operation control. 目標実圧縮比等を示す図である。It is a figure which shows a target actual compression ratio etc. 排気弁の閉弁時期のマップ等を示す図である。It is a figure which shows the map etc. of the valve closing timing of an exhaust valve.

符号の説明Explanation of symbols

1 クランクケース
2 シリンダブロック
3 シリンダヘッド
4 ピストン
5 燃焼室
7 吸気弁
9 排気弁
A 可変圧縮比機構
B 吸気可変バルブタイミング機構
C 排気可変バルブタイミング機構
DESCRIPTION OF SYMBOLS 1 Crankcase 2 Cylinder block 3 Cylinder head 4 Piston 5 Combustion chamber 7 Intake valve 9 Exhaust valve A Variable compression ratio mechanism B Intake variable valve timing mechanism C Exhaust variable valve timing mechanism

Claims (19)

機械圧縮比を変更可能な可変圧縮比機構と、実際の圧縮作用の開始時期を変更可能な実圧縮作用開始時期変更機構と、排気弁とを具備し、機関低負荷運転時には最大の膨張比が得られるように機械圧縮比を最大にし、上記最大の膨張比が20以上であり、更に機関低負荷運転時において排気弁の閉弁時期を吸気上死点の前後10°以内とした、火花点火式内燃機関。 It has a variable compression ratio mechanism that can change the mechanical compression ratio, an actual compression action start timing changing mechanism that can change the actual compression action start time, and an exhaust valve. Spark ignition is performed so that the mechanical compression ratio is maximized so that the maximum expansion ratio is 20 or more and the closing timing of the exhaust valve is within 10 ° before and after the intake top dead center during low engine load operation. Internal combustion engine. 機械圧縮比を変更可能な可変圧縮比機構と、実際の圧縮作用の開始時期を変更可能な実圧縮作用開始時期変更機構と、排気弁の閉弁時期を変更可能な排気可変バルブタイミング機構とを具備し、機関低負荷運転時には最大の膨張比が得られるように機械圧縮比を最大にし、上記最大の膨張比が20以上であり、機関低負荷運転時において排気弁の閉弁時期の設定可能な領域が機関高負荷運転時よりも吸気上死点側に制限された、火花点火式内燃機関。 A variable compression ratio mechanism capable of changing the mechanical compression ratio, an actual compression action start timing changing mechanism capable of changing the actual compression action start timing, and an exhaust variable valve timing mechanism capable of changing the exhaust valve closing timing. provided, it maximizes the mechanical compression ratio so that the maximum expansion ratio is obtained at the time of engine low load operation, and at the maximum expansion ratio is 20 or more, the engine low load during operation of the closing timing of the exhaust valve can be set A spark ignition type internal combustion engine in which the critical region is restricted to the intake top dead center side than during engine high load operation. 機関低負荷運転時において排気弁の閉弁時期を吸気上死点の前後10°以内とした、請求項2に記載の火花点火式内燃機関。 The spark ignition type internal combustion engine according to claim 2, wherein the closing timing of the exhaust valve is within 10 ° before and after the intake top dead center during engine low load operation. 吸気弁の開弁時期を変更可能な吸気可変バルブタイミング機構を更に具備し、機関低負荷運転時に吸気弁の開弁期間と排気弁の開弁期間とが重なるオーバーラップ期間が最小となるように排気弁の閉弁時期及び吸気弁の開弁時期が制御される、請求項2に記載の火花点火式内燃機。   An intake variable valve timing mechanism that can change the opening timing of the intake valve is further provided so that the overlap period in which the intake valve open period and the exhaust valve open period overlap during engine low load operation is minimized. The spark ignition type internal combustion engine according to claim 2, wherein a valve closing timing of the exhaust valve and a valve opening timing of the intake valve are controlled. 吸気弁の開弁時期を変更可能な吸気可変バルブタイミング機構を更に具備し、機関低負荷運転時に吸気弁の開弁期間と排気弁の開弁期間とが重なるオーバーラップ期間がゼロとなるように排気弁の閉弁時期及び吸気弁の開弁時期が制御される、請求項2に記載の火花点火式内燃機関。   An intake variable valve timing mechanism that can change the valve opening timing of the intake valve is further provided so that the overlap period in which the valve opening period of the intake valve overlaps the valve opening period of the exhaust valve becomes zero during low engine load operation. The spark ignition internal combustion engine according to claim 2, wherein the closing timing of the exhaust valve and the opening timing of the intake valve are controlled. 吸気弁の開弁時期を変更可能な吸気弁開弁時期変更機構を更に具備し、機関低負荷運転時には吸気弁の開弁時期を吸気上死点の前後10°以内とした、請求項1又は2に記載の火花点火式内燃機関。 The intake valve opening timing changing mechanism capable of changing the opening timing of the intake valve is further provided, and the opening timing of the intake valve is set within 10 ° before and after the intake top dead center at the time of engine low load operation. 2. The spark ignition internal combustion engine according to 2. 機関低負荷運転時における実圧縮比が機関中高負荷運転時と同じ実圧縮比とされる、請求項1又は2に記載の火花点火式内燃機関。 Engine the actual compression ratio at the time of low load operation is an engine medium and high load operation and the same actual compression ratio, spark-ignition internal combustion engine according to claim 1 or 2. 機関低回転時には機関負荷に関わらずに上記実圧縮比が9〜11の範囲内とされる、請求項7に記載の火花点火式内燃機関。   The spark ignition internal combustion engine according to claim 7, wherein the actual compression ratio is set in a range of 9 to 11 regardless of the engine load at a low engine speed. 機関回転数が高くなるほど上記実圧縮比が高くされる、請求項8に記載の火花点火式内燃機関。   The spark ignition internal combustion engine according to claim 8, wherein the actual compression ratio is increased as the engine speed increases. 上記実圧縮作用開始時期変更機構が吸気弁の閉弁時期を変更可能な吸気可変バルブタイミング機構からなる、請求項1又は2に記載の火花点火式内燃機関。   The spark ignition type internal combustion engine according to claim 1 or 2, wherein the actual compression action start timing changing mechanism comprises an intake variable valve timing mechanism capable of changing a closing timing of the intake valve. 燃焼室内に供給される吸入空気量が吸気弁の閉弁時期を変えることによって制御される、請求項10に記載の火花点火式内燃機関。   The spark ignition internal combustion engine according to claim 10, wherein the intake air amount supplied into the combustion chamber is controlled by changing a closing timing of the intake valve. 吸気弁の閉弁時期は、機関負荷が低くなるにつれて、燃焼室に供給される吸入空気量を制御し得る限界閉弁時期まで吸気下死点から離れる方向に移動せしめられる、請求項11に記載の火花点火式内燃機関。   The valve closing timing of the intake valve is moved in a direction away from the intake bottom dead center until the limit valve closing timing at which the amount of intake air supplied to the combustion chamber can be controlled as the engine load decreases. Spark ignition internal combustion engine. 吸気弁の閉弁時期が上記限界閉弁時期に達したときの機関負荷よりも負荷の高い領域では燃焼室内に供給される吸入空気量が機関吸気通路内に配置されたスロットル弁によらずに吸気弁の閉弁時期を変えることによって制御される、請求項12に記載の火花点火式内燃機関。   In the region where the load of the intake valve is higher than the engine load when the closing timing of the intake valve reaches the above limit closing timing, the amount of intake air supplied into the combustion chamber does not depend on the throttle valve disposed in the engine intake passage. The spark ignition internal combustion engine according to claim 12, which is controlled by changing a closing timing of the intake valve. 吸気弁の閉弁時期が上記限界閉弁時期に達したときの機関負荷よりも負荷の高い領域ではスロットル弁が全開状態に保持される、請求項13に記載の火花点火式内燃機関。   The spark ignition internal combustion engine according to claim 13, wherein the throttle valve is held in a fully open state in a region where the load is higher than the engine load when the closing timing of the intake valve reaches the limit closing timing. 吸気弁の閉弁時期が上記限界閉弁時期に達したときの機関負荷よりも負荷の低い領域では機関吸気通路内に配置されたスロットル弁によって燃焼室内に供給される吸入空気量が制御される、請求項12に記載の火花点火式内燃機関。   In the region where the load is lower than the engine load when the intake valve closing timing reaches the above limit closing timing, the amount of intake air supplied into the combustion chamber is controlled by the throttle valve disposed in the engine intake passage. The spark ignition internal combustion engine according to claim 12. 吸気弁の閉弁時期が上記限界閉弁時期に達したときの機関負荷よりも負荷の低い領域では負荷が低くなるほど空燃比が大きくされる、請求項12に記載の火花点火式内燃機関。   The spark ignition internal combustion engine according to claim 12, wherein the air-fuel ratio is increased as the load decreases in a region where the load is lower than the engine load when the closing timing of the intake valve reaches the limit closing timing. 吸気弁の閉弁時期が上記限界閉弁時期に達したときの機関負荷よりも負荷の低い領域では吸気弁の閉弁時期が上記限界閉弁時期に保持される、請求項12に記載の火花点火式内燃機関。   The spark according to claim 12, wherein the closing timing of the intake valve is held at the limit closing timing in a region where the load is lower than the engine load when the closing timing of the intake valve reaches the limit closing timing. Ignition internal combustion engine. 上記機械圧縮比は機関負荷が低くなるにつれて限界機械圧縮比まで増大せしめられる、請求項1又は2に記載の火花点火式内燃機関。   The spark ignition internal combustion engine according to claim 1 or 2, wherein the mechanical compression ratio is increased to a limit mechanical compression ratio as the engine load decreases. 上記機械圧縮比が上記限界機械圧縮比に達したときの機関負荷よりも負荷の低い領域では機械圧縮比が上記限界機械圧縮比に保持される、請求項18に記載の火花点火式内燃機関。   19. The spark ignition internal combustion engine according to claim 18, wherein the mechanical compression ratio is maintained at the limit mechanical compression ratio in a region where the load is lower than the engine load when the mechanical compression ratio reaches the limit mechanical compression ratio.
JP2006192832A 2006-07-13 2006-07-13 Spark ignition internal combustion engine Expired - Fee Related JP4259546B2 (en)

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